260

Click here to load reader

Centrifugal Pumps Handbook

Embed Size (px)

DESCRIPTION

CF pump basics

Citation preview

Page 1: Centrifugal Pumps Handbook

The Pump Handbook Series 1

apor pressure, cavitation,and NPSH are subjectswidely discussed by engi-neers, pumps users, and

pumping equipment suppliers, butunderstood by too few. To graspthese subjects, a basic explanationis required.

VAPOR PRESSUREKnowledge of vapor pressure

is extremely important whenselecting pumps and theirmechanical seals. Vapor pressureis the pressure absolute at which aliquid, at a given temperature,starts to boil or flash to a gas.Absolute pressure (psia) equals thegauge pressure (psig) plus atmos-pheric pressure.

Let’s compare boiling water atsea level in Rhode Island to boil-ing water at an elevation of 14,110feet on top of Pikes Peak inColorado. Water boils at a lowertemperature at altitude becausethe atmospheric pressure is lower.

Water and water containingdissolved air will boil at differenttemperatures. This is because oneis a liquid and the other is a solu-tion. A solution is a liquid with dis-solved air or other gases. Solutionshave a higher vapor pressure than

their parent liquid and boil at a lowertemperature. While vapor pressurecurves are readily available for liq-uids, they are not for solutions.Obtaining the correct vapor pressurefor a solution often requires actuallaboratory testing.

CAVITATIONCavitation can create havoc with

pumps and pumping systems in theform of vibration and noise. Bearingfailure, shaft breakage, pitting on theimpeller, and mechanical seal leak-age are some of the problems causedby cavitation.

When a liquid boils in the suc-tion line or suction nozzle of a pump,it is said to be “flashing” or “cavitat-ing” (forming cavities of gas in theliquid). This occurs when the pres-sure acting on the liquid is below thevapor pressure of the liquid. Thedamage occurs when these cavitiesor bubbles pass to a higher pressureregion of the pump, usually just pastthe vane tips at the impeller “eye,”and then collapse or “implode.”

NPSHNet Positive Suction Head is the

difference between suction pressureand vapor pressure. In pump designand application jargon, NPSHA is the

net positive suctionhead available to thepump, and NPSHR isthe net positive suc-tion head requiredby the pump.

The NPSHAmust be equal to orgreater than theNPSHR for a pumpto run properly. Oneway to determine theNPSHA is to mea-sure the suction pres-sure at the suctionnozzle, then applythe following formu-la:

NPSHA = PB – VP ± Gr+ hv

where PB = barometric pres-sure in feet absolute, VP = vaporpressure of the liquid at maximumpumping temperature in feetabsolute, Gr = gauge reading atthe pump suction, in feet absolute(plus if the reading is above baro-metric pressure, minus if the read-ing is below the barometricpressure), and hv = velocity headin the suction pipe in feetabsolute.

NPSHR can only be deter-mined during pump testing. Todetermine it, the test engineermust reduce the NPSHA to thepump at a given capacity until thepump cavitates. At this point thevibration levels on the pump andsystem rise, and it sounds likegravel is being pumped. Morethan one engineer has run for theemergency shut-down switch thefirst time he heard cavitation onthe test floor. It’s during thesetests that one gains a real apprecia-tion for the damage that will occurif a pump is allowed to cavitate fora prolonged period.

CENTRIFUGAL PUMPINGCentrifugal pumping terminol-

ogy can be confusing. The follow-ing section addresses these termsand how they are used:

Head is a term used toexpress pressure in both pumpdesign and system design whenanalyzing static or dynamic condi-tions. This relationship isexpressed as:

head in feet = (pressure in psi x 2.31)

specific gravity

Pressure in static systems isreferred to as static head and in adynamic system as dynamichead.

To explain static head, let’sconsider three columns of anydiameter, one filled with water,one with gasoline, and one withsalt water (Figure 1). If thecolumns are 100 ft tall and you

Nomenclature and DefinitionsBY PAT FLACH

V

FIGURE 1

Static head using various liquids.

43 psi 52 psi32.5 psi

CENTRIFUGAL PUMPS

HANDBOOK

100FEET

STATICHEAD

100FEET

STATICHEAD

100FEET

STATICHEAD

WaterSp. Gr. = 1.0

GasolineSp. Gr. = .75

SaltWaterSp. Gr. = 1.2

Page 2: Centrifugal Pumps Handbook

2 The Pump Handbook Series

measure the pressure at the bot-tom of each column, the pres-sures would be 43, 32.5, and 52psi, respectively. This is becauseof the different specific gravities,or weights, of the three liquids.Remember, we are measuringpounds per square inch at thebottom of the column, not thetotal weight of the liquid in thecolumn.

The following four terms areused in defining pumping systemsand are illustrated in Figure 2.

Total static head is the verti-cal distance between the surfaceof the suction source liquid andthe surface level of the dischargeliquid.

Static discharge head is thevertical distance from the center-line of the suction nozzle up tothe surface level of the dischargeliquid.

Static suction head applieswhen the supply is above thepump. It is the vertical distancefrom the centerline of the suctionnozzle up to the liquid surface ofthe suction supply.

Static suction lift applieswhen the supply is located belowthe pump. It is the vertical dis-tance from the centerline of thesuction nozzle down to the surfaceof the suction supply liquid.

Velocity, friction, and pressurehead are used in conjunction withstatic heads to define dynamicheads.

Velocity head is the energy ina liquid as a result of it traveling atsome velocity V. It can be thoughtof as the vertical distance a liquidwould need to fall to gain the samevelocity as a liquid traveling in apipe.This relationship is expressed as:

hv = V2/2g

where V = velocity of theliquid in feet per second and g =32.2 ft/sec2.

Friction head is the headneeded to overcome resistance toliquid flowing in a system. This

at a pump suction flange, convert-ing it to head and correcting to thepump centerline, then adding thevelocity head at the point of thegauge.

Total dynamic dischargehead is the static discharge headplus the velocity head at the pumpdischarge flange plus the total fric-tion head in the discharge system.This can be determined in the fieldby taking the discharge pressurereading, converting it to head, andcorrecting it to the pump center-line, then adding the velocityhead.

Total dynamic suction lift isthe static suction lift minus thevelocity head at the suction flangeplus the total friction head in thesuction line. To calculate totaldynamic suction lift, take suctionpressure at the pump suctionflange, convert it to head and cor-rect it to the pump centerline, thensubtract the velocity head at thepoint of the gauge.

Total dynamic head in asystem is the total dynamic dis-charge head minus the totaldynamic suction head when thesuction supply is above the pump.When the suction supply is belowthe pump, the total dynamic head

resistance can come from pipe fric-tion, valves, and fittings. Values infeet of liquid can be found in theHydraulic Institute Pipe FrictionManual.

Pressure head is the pressure infeet of liquid in a tank or vessel on thesuction or discharge side of a pump. Itis important to convert this pressureinto feet of liquid when analyzing sys-tems so that all units are the same. If avacuum exists and the value is knownin inches of mercury, the equivalentfeet of liquid can be calculated usingthe following formula:

vacuum in feet = in. of Hg x 1.13specific gravity

When discussing how a pumpperforms in service, we use termsdescribing dynamic head. In otherwords, when a pump is running it isdynamic. Pumping systems are alsodynamic when liquid is flowingthrough them, and they must be ana-lyzed as such. To do this, the follow-ing four dynamic terms are used.

Total dynamic suction head isthe static suction head plus the veloc-ity head at the suction flange minusthe total friction head in the suctionline. Total dynamic suction head iscalculated by taking suction pressure

FIGURE 2

Total static head, static discharge head, static suction head, and static suction lift.

TotalStaticHead

StaticDischarge

Head

StaticSuctionHead

StaticDischargeHead

TotalStaticHead

StaticSuction

Lift

Page 3: Centrifugal Pumps Handbook

The Pump Handbook Series 3

is the total dynamic discharge headplus the total dynamic suction lift.

Centrifugal pumps are dynamicmachines that impart energy to liq-uids. This energy is imparted bychanging the velocity of the liquid asit passes through the impeller. Mostof this velocity energy is then con-verted into pressure energy (totaldynamic head) as the liquid passesthrough the casing or diffuser.

To predict the approximate totaldynamic head of any centrifugalpump, we must go through two steps.First, the velocity at the outside diam-eter (o.d.) of the impeller is calculatedusing the following formula:

v = (rpm x D)/229

where v = velocity at the periph-ery of the impeller in ft per second, D= o.d. of the impeller in inches, rpm= revolutions per minute of theimpeller, and 229 = a constant.

Second, because the velocityenergy at the o.d. or periphery of theimpeller is approximately equal to thetotal dynamic head developed by thepump, we continue by substituting vfrom above into the following equa-tion:

H = v2/2g

where H = total dynamic headdeveloped in ft, v = velocity at theo.d. of the impeller in ft/sec, and g =32.2 ft/sec2.

A centrifugal pump operating ata given speed and impeller diameterwill raise liquid of any specific gravi-ty or weight to a given height.Therefore, we always think in termsof feet of liquid rather than pressurewhen analyzing centrifugal pumpsand their systems. ■

Patrick M. Flach is the westernhemisphere Technical Services Managerfor the Industrial Division of EG&GSealol.

Have you had a momentary (or continuing) problem with con-verting gallons per minute to cubic meters per second or liters persecond? Join the crowd. Though the metric or SI system is probablyused as the accepted system, more than English units, it still presentsa problem to a lot of engineers.

Authors are encouraged to use the English system. Following is alist of the common conversions from English to metric units. This isfar from a complete list. It has been limited to conversions frequentlyfound in solving hydraulic engineering problems as they relate topumping systems.

PUMPING UNITS

FLOW RATE(U.S.) gallons/min (gpm) x 3.785 = liters/min (L/min)(U.S.) gpm x 0.003785 = cubic meters/min (m3/min)cubic feet/sec (cfs) x 0.028 = cubic meters/sec (m3/s)

HEADfeet (ft) x 0.3048 = meters (m)pounds/square inch (psi) x 6,895 = Pascals (Pa)

POWERhorsepower (Hp) x 0.746 = kilowatts (kW)

GRAVITATIONAL CONSTANT (g)32.2 ft./s2 x 0.3048 = 9.81 meters/second2 (m/s2)

SPECIFIC WEIGHTlb/ft3 x 16.02 = kilogram/cubic meter (kg/m3)

VELOCITY (V)ft/s x 0.3048 = meters/second (m/s)

VELOCITY HEADV2/2g (ft) x 0.3048 = meters (m)

SPECIFIC SPEED (Ns)(gpm–ft) x 0.15 = Ns(m3/min–m)Ns = N(rpm)[(gpm)0.5/(ft)0.75]

J. Robert Krebs is President of Krebs Consulting Service. He serves onthe Pumps and Systems Editorial Advisory Board.

Basic Units Multiply English x Factor = MetricLength Feet x 0.3048 = Meter (m)Mass Pound x 0.454 = Kilogram (Kg)Force Pound x 4.448 = Newton (N)Pressure Pound/Square In. (psi) x 6,895 = Pascal (Pa)Time Seconds x 1 = Seconds (s)Gallon (US) Gallon x 0.003785 = Meter Cubed (m3)Gallon (US) Gallon x 3.785 = Liter (L)

TABLE 1. ENGLISH TO METRIC CONVERSION

Pumping Terms

Page 4: Centrifugal Pumps Handbook

4 The Pump Handbook Series

Centrifugal and Positive Displacement Pumps in the Operating System

n the many differences that existbetween centrifugal and positivedisplacement pumps, one whichhas caused some confusion is the

manner in which they each operatewithin the system.

Positive displacement pumps havea series of working cycles, each ofwhich encloses a certain volume offluid and moves it mechanicallythrough the pump into the system,regardless of the back pressure on thepump. While the maximum pressuredeveloped is limited only by themechanical strength of the pump andsystem and by the driving poweravailable, the effect of that pressurecan be controlled by a pressure reliefor safety valve.

A major advantage of the posi-tive displacement pump is that itcan deliver consistent capacitiesbecause the output is solely depen-dent on the basic design of thepump and the speed of its drivingmechanism. This means that, if therequired flow rate is not movingthrough the system, the situationcan always be corrected by chang-ing one or both of these factors.

This is not the case with the cen-trifugal pump, which can onlyreact to the pressure demand of thesystem. If the back pressure on acentrifugal pump changes, so willits capacity.

This can be disruptive for anyprocess dependent on a specificflow rate, and it can diminish theoperational stability, hydraulic effi-ciency and mechanical reliability ofthe pump.

CENTRIFUGAL PUMP PERFORMANCE CURVE

The interdependency of the sys-tem and the centrifugal pump can beeasily explained with the use of thepump performance curve and the system curve.

A centrifugal pump performancecurve is a well known shape whichshows that the pressure the pump

can develop is reduced as the capacityincreases. Conversely, as the capacitydrops, the pressure it can achieve isgradually increased until it reaches amaximum where no liquid can passthrough the pump. Since this is usuallya relatively low pressure, it is rarelynecessary to install a pressure relief orsafety valve.

When discussing the pressuresdeveloped by a centrifugal pump, weuse the equivalent linear measurementreferred to as “head,” which allows thepump curve to apply equally to liquids of different densities.

[Head (in feet)=Pressure (in p.s.i.) x2.31+ Specific Gravity of the liquid]

SYSTEM CURVEThe system curve represents the

pressures needed at different flow ratesto move the product through the sys-tem. To simplify a comparison withthe centrifugal pump curve, we againuse the ‘head’ measurement. The sys-tem head consists of three factors:• static head, or the vertical eleva-

tion through which the liquidmust be lifted

• friction head, or the head requiredto overcome the friction losses inthe pipe, the valves and all the fit-tings and equipment

• velocity head, which is the headrequired to accelerate the flow ofliquid through the pump (Velocityhead is generally quite small andoften ignored.)

As the static head does not varysimply because of a change in flowrate, the graph would show a straightline. However, both the friction head and thevelocity headwill alwaysvary direct- ly with thecapacity. Thecombinationof all threecreates thesystem curve.

When the pump curve is super-imposed on the system curve, thepoint of intersection represents theconditions (H,Q) at which the pumpwill operate.

Pumping conditions changeONLY through an alteration ineither the pump curve or the sys-tem curve.

When considering possiblemovements in these curves, itshould be noted that there are onlya few conditions which will causethe pump curve to change its posi-tion and shape:

• wear of the impeller• change in rotational speed• change in impeller diameter• change in liquid viscosity Since these conditions don’t nor-

mally develop quickly, any suddenchange in pumping conditions islikely to be a result of a movementin the system curve, which meanssomething in the system haschanged.

Since there are only three ingre-dients in a system curve, one ofwhich is minimal, it follows thateither the static head or the frictionhead must have changed for anymovement to take place in the sys-tem curve.

A change in the statichead is normally a result ofa change in tank level. Ifthe pump is emptying atank and discharging at afixed elevation, the statichead against which thepump must operate will begradually increasing as the

IBY ROSS C. MACKAY

Head

Capacity

SystemCurve Friction &

Velocity Head

Static Head

Pump Curve

System Curve

H

OQ

CENTRIFUGAL PUMPS

HANDBOOK

Page 5: Centrifugal Pumps Handbook

The Pump Handbook Series 5

suction tank empties. This will causethe system curve to move upwardsas shown.

An increase in friction head canbe caused by a wide variety of con-ditions such as the change in a valvesetting or build-up of solids in astrainer. This will give the systemcurve a new slope.

Both sets of events produce thesame result: a reduction of flowthrough the system. If the flow isredirected to a different location(such as in a tank farm), it meansthat the pump is now operating onan entirely new system which willhave a completely different curve.

Thus, it is clear that regardless ofthe rated capacity of the centrifugalpump, it will only provide what thesystem requires. It is important tounderstand the conditions underwhich system changes occur, theacceptability of the new operatingpoint on the pump curve, and themanner in which it can be moved.

When the operating conditions of asystem fitted with a centrifugal pumpchange, it is helpful to consider thesecurves, focus on how the system iscontrolling the operation of the pump,and then control the system in theappropriate way. ■

Ross C. Mackay is an independent con-sultant located in Tottenham, Ontario,Canada. He is the author of several paperson the practical aspects of pump mainte-nance, and a specialist in helping companiesreduce their pump maintenance costs.

Page 6: Centrifugal Pumps Handbook

6 The Pump Handbook Series

avitation is the formationand collapse of vapor bub-bles in a liquid.

Bubble formationoccurs at a point where the pres-sure is less than the vapor pres-sure, and bubble collapse orimplosion occurs at a point wherethe pressure is increased to thevapor pressure. Figure 1 showsvapor pressure temperature char-acteristics.

This phenomenon can alsooccur with ship propellers and inother hydraulic systems such asbypass orifices and throttlevalves—situations where anincrease in velocity with resultingdecrease in pressure can reducepressure below the liquid vaporpressure.

CAVITATION EFFECTS

BUBBLE FORMATION PHASEFlow is reduced as the liquid

is displaced by vapor, andmechanical imbalance occurs asthe impeller passages are partiallyfilled with lighter vapors. Thisresults in vibration and shaftdeflection, eventually resulting inbearing failures, packing or sealleakage, and shaft breakage. Inmulti-stage pumps this can causeloss of thrust balance and thrustbearing failures.

BUBBLE COLLAPSE PHASE1. Mechanical damage occurs as

the imploding bubbles removesegments of impeller material.

2. Noise and vibration result fromthe implosion. Noise thatsounds like gravel beingpumped is often the user’s firstwarning of cavitation.

NET POSITIVE SUCTION HEADWhen designing a pumping

system and selecting a pump, onemust thoroughly evaluate net posi-tive suction head (NPSH) to pre-vent cavitation. A proper analysis

Cavitation and NPSH in Centrifugal PumpsBY PAUL T. LAHR

C

FIGURE 1

Vapor pressures of various liquids related to temperature.

involves both the net positive suctionheads available in the system(NPSHA) and the net positive suctionhead required by the pump (NPSHR).

NPSHA is the measurement orcalculation of the absolute pressureabove the vapor pressure at thepump suction flange. Figure 2 illus-trates methods of calculating NPSHAfor various suction systems. Since

friction in the suction pipe is acommon negative component ofNPSHA, the value of NPSHA willalways decrease with flow.

NPSHA must be calculated toa stated reference elevation, suchas the foundation on which thepump is to be mounted.

NPSHR is always referencedto the pump impeller center line.

CENTRIFUGAL PUMPS

HANDBOOK

1000

800

600500

400

300

200

100

80

605040

30

20

10

8

654

3

2

1.0

.80

.60

.50

.40

.30

.20

.1060 30 0 30 60 90 120 150 180 210 240

28"

28.5"

29"29.1"29.2"29.3"

29.4"

29.5"

29.6"

29.7"29.72"

10"

15"

20"

22.5"

25"

26"

27"

80

60

50

40

30

20

14

1052 05

985

800

600500

400

300

200

140

100

-60° to 240°F

TEMPERATURE–F

CARBON DIOXIDE

NITROUS OXIDE

ETHANE

MONOCHLOROTRIFLUOROMETHANE

HYDROGEN SULFIDE

PROPYLENE

PROPANE

AMMONIA

CHLORINE

METHYL CHLORID

ESULFUR D

IOXID

E

ISOBUTANE BUTANE

ETHYL CHLO

RIDE

METHYL FORMATE

DIETHYL E

THER

METHYLENE C

HLORID

E

DICHLO

ROETHYLENE

ACETONE

DICHLO

ROETH

YLENE (C

IS)

CHLOROFORM (TRIC

HLOROMETHANE)

CARBON TETRACHLO

RIDE TRIC

HLOROETHULENE

WATE

R

HEAVYWATER (S

P.GR. A

T 70 F

=1.106

GA

UG

E P

RE

SS

UR

E–L

BS

. PE

R S

Q. I

N.

VA

CU

UM

–IN

CH

ES

OF

ME

RC

UR

Y

AB

SO

LUT

E P

RE

SS

UR

E–L

BS

. PE

R S

Q. I

N.

Page 7: Centrifugal Pumps Handbook

The Pump Handbook Series 7

It is a measure of the pressuredrop as the liquid travels fromthe pump suction flange along theinlet to the pump impeller. Thisloss is due primarily to frictionand turbulence.

Turbulence loss is extremelyhigh at low flow and thendecreases with flow to the bestefficiency point. Friction lossincreases with increased flow. Asa result, the internal pump losseswill be high at low flow, drop-ping at generally 20–30% of thebest efficiency flow, then increas-ing with flow. The complex sub-ject of turbulence and NPSHR atlow flow is best left to anotherdiscussion.

Figure 3 shows the pressureprofile across a typical pump at afixed flow condition. The pres-sure decrease from point B topoint D is the NPSHR for thepump at the stated flow.

The pump manufacturerdetermines the actual NPSHR foreach pump over its completeoperating range by a series oftests. The detailed test procedureis described in the HydraulicInstitute Test Standard 1988Centrifugal Pumps 1.6. Industryhas agreed on a 3% head reduc-tion at constant flow as the stan-dard value to establish NPSHR.Figure 4 shows typical results of aseries of NPSHR tests.

The pump system designermust understand that the pub-lished NPSHR data establishedabove are based on a 3% headreduction. Under these condi-tions the pump is cavitating. Atthe normal operating point theNPSHA must exceed the NPSHRby a sufficient margin to elimi-nate the 3% head drop and theresulting cavitation.

The NPSHA margin requiredwill vary with pump design andother factors, and the exact mar-gin cannot be precisely predicted.For most applications the NPSHAwill exceed the NPSHR by a sig-nificant amount, and the NPSHmargin is not a consideration. Forthose applications where theNPSHA is close to the NPSHR

FIGURE 2

FIGURE 3

A B C D E

The pressure profile across a typical pump at a fixed flow condition.

Calculation of system net positive suction head available (NPSHA) for typicalsuction conditions. PB = barometric pressure in feet absolute, VP = vaporpressure of the liquid at maximum pumping temperature in feet absolute, p =pressure on surface of liquid in closed suction tank in feet absolute, Ls = max-imum suction lift in feet, LH = minimum static suction head in feet, hf = fric-tion loss in feet in suction pipe at required capacity.

E

A B C

D

ENTRANCELOSS

FRICTION

TURBULANCEFRICTION

ENTRANCELOSS AT

VANE TIPS

INCREASINGPRESSURE

DUE TO IMPELLER

PO

INT

OF

LOW

ES

T P

RE

SS

UR

EW

HE

RE

VA

PO

RIZ

AT

ION

ST

AR

TS

INC

RE

AS

E P

RE

SS

UR

E

4a SUCTION SUPPLY OPEN TO ATMOSPHERE-with Suction Lift

CL

PBNPSHA=PB – (VP + Ls + ht)

4b SUCTION SUPPLY OPEN TO ATMOSPHERE-with Suction Head

NPSHA=PB + LH – (VP + ht)

PB

CL

4c CLOSED SUCTION SUPPLY -with Suction Lift

NPSHA=p – (Ls + VP + ht)

p

CL

4d CLOSED SUCTION SUPPLY -with Suction Lift

NPSHA=p + LH – (VP + ht)

p

CL

POINTS ALONG LIQUID PATHRELATIVE PRESSURES IN THE ENTRANCE SECTION OF A PUMP

Page 8: Centrifugal Pumps Handbook

8 The Pump Handbook Series

(2–3 feet), users should consult thepump manufacturer and the twoshould agree on a suitable NPSHmargin. In these deliberations, fac-tors such as liquid characteristic,minimum and normal NPSHA,and normal operating flow mustbe considered.

SUCTION SPECIFIC SPEEDThe concept of suction specif-

ic speed (Ss) must be consideredby the pump designer, pumpapplication engineer, and the sys-tem designer to ensure a cavita-tion-free pump with highreliability and the ability to oper-ate over a wide flow range.

N x Q0.5Ss = ——————

(NPSHR)0.75

where N = pump rpmQ = flow rate in gpm at the

best efficiency pointNPSHR = NPSHR at Q with

the maximum impeller diameter

The system designer shouldalso calculate the system suction

specific speed by substi-tuting design flow rate andthe system designer’sNPSHA. The pump speedN is generally determinedby the head or pressurerequired in the system.For a low-maintenancepump system, designersand most user specifica-tions require, or prefer, Ssvalues below 10,000 to12,000. However, as indi-cated above, the pump Ssis dictated to a greatextent by the system con-ditions, design flow, head,and the NPSHA.

Figures 5 and 6 areplots of Ss versus flow ingpm for various NPSHAor NPSHR at 3,500 and1,750 rpm. Similar plotscan be made for other commonpump speeds.

Using curves from Figure 5 andFigure 6 allows the system designerto design the system Ss, i.e., for a sys-tem requiring a 3,500 rpm pumpwith 20 feet of NPSHA, the maxi-mum flow must be limited to 1,000

gpm if the maximum Ss is to bemaintained at 12,000. Variousoptions are available, such asreducing the head to allow 1,750rpm (Figure 7). This would allowflows to 4,000 gpm with 20 feet ofNPSHA.

Q1Q2

100% CAP Q3

Q4

NPSH

3%

NP

SH

R

TO

TA

L H

EA

D

FIGURE 4

Typical results of a four-point net posi-tive suction head required (NPSHR) testbased on a 3% head drop.

1 2 3 4 5 6 7 8 9 1 2 3 4 5 6 7 8 9 1 2 3 4 5 6 7 8 9 1

3

2

198

7

6

5

4

3

2

1

HSV=2

3

4

5

678

910

HSV=12

14 16 18 20

28

32

3640

50 55

60 65

HSV=24

HSV=45

Solution for

S=N

for N=3,500 rpm

Q Hsv

0.75

A plot of suction specific speed (Ss) versus flow in gallons per minute (gpm) for various NPSHA orNPSHR at 3,500 rpm. (Single suction pumps. For double suction use 1/2 capacity). Hsv=NPSHR atBEP with maximum impeller diameter.

Q, Capacity, gpm

S, S

uctio

n sp

ecifi

c sp

eed

FIGURE 5

Page 9: Centrifugal Pumps Handbook

The Pump Handbook Series 9

It is important forthe pump user to under-stand how critical thesystem design require-ments are to the selec-tion of a reliable,trouble-free pump.

Matching the systemand pump characteristicsis a must. Frequently,more attention is paid tothe discharge side. Yet itis well known that mostpump performanceproblems are causedby problems on thesuction side.

Figure 7 is a typicalplot of the suction anddischarge systems.

It is important thatpoints A, B, and C be wellestablished and under-stood. A is the normaloperating point. B is themaximum flow for cavi-tation-free operation. C isthe minimum stable flow,which is dictated by thesuction specific speed.

As a general rule, the higherthe suction specific speed, thehigher the minimum stable flowcapacity will be. If a pump isalways operated at its best efficien-cy point, a high value of Ss will notcreate problems. However, if thepump is to be operated at reducedflow, then the Ss value must begiven careful consideration. ■

REFERENCES1. Goulds Pump Manual.

2. Durco Pump Engineering Manual.

3. Hydraulic Institute TestStandards—1988 CentrifugalPumps 1.6.

Paul T. Lahr is the owner ofPump Technology, a consulting firm.He serves on the Pumps andSystems Editorial Advisory Board.

HE

AD

NP

SH

- F

EE

T

GPMC A B

4

3

2

1

FIGURE 7

1 2 3 4 5 6 7 8 9 1 2 3 4 5 6 7 8 9 1 2 3 4 5 6 7 8 9 1

5

4

3

2

198

7

6

5

4

3

2

1

HSV=12

HSV=1

1098

6

3

2

14

45

7 1618 20

2832 36 40

50

HSV=24

HSV=45

Solution for

S=N

for N=1,750 rpm

Q Hsv

0.75

A plot of suction specific speed (Ss) versus flow in gallons per minute (gpm) for various NPSHA orNPSHR at 1,750 rpm. (Single suction pumps. For double suction use 1/2 capacity.) HSV=NPSHR atBEP with the maximum impeller diameter.

A typical plot of the suction and dischargesystems. Curve 1 = pump head capacityperformance, curve 2 = total system curve,curve 3 = suction system curve NPSHA,and curve 4 = pump NPSHR.

Q, Capacity, gpm

S, S

uctio

n sp

ecifi

c sp

eed

FIGURE 6

Page 10: Centrifugal Pumps Handbook

10 The Pump Handbook Series

f a wide receiver has the rightspeed and good hands, all that’sneeded from the quarterback isto throw the ball accurately,

and the team will probably gaingood yardage, maybe even atouchdown.

Believe it or not, much thesame is true of a pump and its suc-tion conditions. If it has the rightspeed and is the right size, allthat’s required from the quarter-back is to deliver the liquid at theright pressure and with an evenlaminar flow into the eye of theimpeller.

If the quarterback’s pass is offtarget, badly timed, or the ball’sturning end over end in the air,the receiver may not be able tohang on to it, and there’s no gainon the play. When that hap-pens, the quarterbackknows he didn’t throw itproperly and doesn’t blamethe receiver. Unfortunately,that’s where the compari-son ends. The engineering”quarterbacks” tend toblame the pump even whenits their delivery that’s bad!

Just as there are tech-niques a quarterback mustlearn in order to throwaccurately, there are ruleswhich ensure that a liquidarrives at the impeller eye withthe pressure and flow characteris-tics needed for reliable operation.

RULE #1. PROVIDE SUFFICIENT NPSH

Without getting too complicat-ed on a subject about which com-plete books have been written,let’s just accept the premise thatevery impeller requires a mini-mum amount of pressure energyin the liquid being supplied inorder to perform without cavita-tion difficulties. This pressureenergy is referred to as NetPositive Suction Head Required.

The NPSH Available is sup-plied from the system. It is solely

a function of the system design onthe suction side of the pump.Consequently, it is in the control ofthe system designer.

To avoid cavitation, the NPSHavailable from the system must begreater than the NPSH required bythe pump, and the biggest mistakethat can be made by a system design-er is to succumb to the temptation toprovide only the minimum requiredat the rated design point. This leavesno margin for error on the part of thedesigner, or the pump, or the system.Giving in to this temptation hasproved to be a costly mistake onmany occasions.

In the simple system as shownin Figure 1, the NPSH Available canbe calculated as follows:

NPSHA = Ha + Hs - Hvp - Hf

where Ha= the head on the surface of the

liquid in the tank. In an opensystem like this, it will beatmospheric pressure.

Hs= the vertical distance of thefree surface of the liquidabove the center line of thepump impeller. If the liquid isbelow the pump, thisbecomes a negative value.

Hvp= the vapor pressure of the liq-uid at the pumping tempera-ture, expressed in feet ofhead.

Hf= the friction losses in thesuction piping.

The NPSH Available may alsobe determined with this equation:

NPSHA= Ha + Hg + V2/2g - Hvp

whereHa= atmospheric pressure in

feet of head.Hg= the gauge pressure at the

suction flange in feet ofhead.

V2= The velocity head at thepoint of measurement ofHg. (Gauge readings do notinclude velocity head.)

RULE #2. REDUCE THE FRICTION LOSSES

When a pump is taking itssuction from a tank, it should belocated as close to the tank as pos-sible in order to reduce the effectof friction losses on the NPSHAvailable. Yet the pump must befar enough away from the tank toensure that correct piping practicecan be followed. Pipe friction canusually be reduced by using a larg-er diameter line to limit the linearvelocity to a level appropriate tothe particular liquid beingpumped. Many industries workwith a maximum velocity of about5ft./sec., but this is not alwaysacceptable.

RULE #3.NO ELBOWS ON THE SUCTION FLANGE

Much discussion has takenplace over the acceptable configu-ration of an elbow on the suctionflange of a pump. Let’s simplify it.There isn’t one!

There is always an unevenflow in an elbow, and when one isinstalled on the suction of anypump, it introduces that unevenflow into the eye of the impeller.This can create turbulence and air

Pump Suction ConditionsBY ROSS C. MACKAY

I

FIGURE 1

2g

CENTRIFUGAL PUMPS

HANDBOOK

Ha

Hvp

Hf

Hs

Page 11: Centrifugal Pumps Handbook

The Pump Handbook Series 11

entrainment, which may result inimpeller damage and vibration.

When the elbow is installedin a horizontal plane on the inletof a double suction pump,uneven flows are introduced intothe opposing eyes of theimpeller, upsetting the hydraulicbalance of the rotating element.Under these conditions the over-loaded bearing will fail prema-turely and regularly if the pumpis packed. If the pump is fittedwith mechanical seals, the sealwill usually fail instead of thebearing-but just as regularly andoften more frequently.

The only thing worse thanone elbow on the suction of apump is two elbows on the suc-tion of a pump— particularly ifthey are positioned in planes atright angles to each other. Thiscreates a spinning effect in theliquid which is carried into theimpeller and causes turbulence,inefficiency and vibration.

A well established and effec-tive method of ensuring a lami-nar flow to the eye of theimpeller is to provide the suctionof the pump with a straight run

of pipe in a lengthequivalent to 5-10times the diameterof that pipe. Thesmaller multiplierwould be used onthe larger pipediameters and viceversa.

RULE #4. STOP AIROR VAPOR ENTERINGTHE SUCTION LINE

Any high spotin the suction linecan become filledwith air or vapor which, if trans-ported into the impeller, will createan effect similar to cavitation andwith the same results. Serviceswhich are particularly susceptibleto this situation are those where thepumpage contains a significantamount of entrained air or vapor,as well as those operating on a suc-tion lift, where it can also cause thepump to lose its prime. (Figure 3)

A similar effect can becaused by a concentricreducer. The suction of apump should be fitted withan eccentric reducer posi-

tioned withthe flat sideuppermos t .(Figure 4).

If a pumpis taking itssuction froma sump ortank, the for-mation of vortices candraw air into the suc-tion line. This can usu-ally be prevented byproviding sufficientsubmergence of liquidover the suction open-ing. A bell-mouth designon the opening willreduce the amount ofsubmergence required.This submergence iscompletely independentof the NPSH required bythe pump.

It is worthwhilenoting that these vor-

tices are more difficult to trou-bleshoot in a closed tank simplybecause they can’t be seen aseasily.

Great care should be takenin designing a sump to ensurethat any liquid emptying into itdoes so in such a way that airentrained in the inflow does notpass into the suction opening.Any problem of this nature may

require a change in the relativepositions of the inflow and outletif the sump is large enough, orthe use of baffles. (Figure 5)

RULE #5. CORRECT PIPING ALIGNMENT

Piping flanges must be accu-rately aligned before the boltsare tightened and all piping,valves and associated fittingsshould be independently sup-ported, so as to place no strainon the pump. Stress imposed onthe pump casing by the pipingreduces the probability of satis-factory performance.

FIGURE 2

FIGURE 4

FIGURE 3

Air Pocket

Suction

Page 12: Centrifugal Pumps Handbook

12 The Pump Handbook Series

Under certain conditions thepump manufacturer may identifysome maximum levels of forcesand moments which may beacceptable on the pump flanges.

In high temperature applica-tions, some piping misalignment is inevitable owing to thermalgrowth during the operating cycle.Under these conditions, thermalexpansion joints are often intro-duced to avoid transmitting pipingstrains to the pump. However, ifthe end of the expansion jointclosest to the pump is notanchored securely, the object ofthe exercise is defeated as the pip-ing strains are simply passedthrough to the pump.

RULE #6. WHEN RULES 1 TO5 HAVE BEENIGNORED, FOLLOWRULES 1 TO 5.

Piping designis one area wherethe basic princi-ples in-volved areregularly ignored,resulting inhydraulic instabil-

ities in the impeller which trans-late into additional shaft loading,higher vibration levels and pre-mature failure of the seal or bear-ings. Because there are manyother reasons why pumps couldvibrate, and why seals and bear-ings fail, the trouble is rarelytraced to incorrect piping.

It has been argued thatbecause many pumps are pipedincorrectly and most of them areoperating quite satisfactorily, pip-ing procedure is not important.Unfortunately, satisfactory opera-tion is a relative term, and whatmay be acceptable in one plantmay be inappropriate in another.

Even when ”satisfactory”pump operation is obtained, that

doesn’t automatically make aquestionable piping practice cor-rect. It merely makes it lucky.

The suction side of a pump ismuch more important than thepiping on the discharge. If anymistakes are made on the dis-charge side, they can usually becompensated for by increasingthe performance capability fromthe pump. Problems on the suc-tion side, however, can be thesource of ongoing and expensivedifficulties which may never betraced back to that area.

In other words, if yourreceivers aren’t performing well,is it their fault? Or does the quar-terback need more training? ■

Ross C. Mackay is an indepen-dent consultant who specializes inadvanced technology training forpump maintenance cost reduction.He also serves on the editorial adviso-ry board for Pumps and Systems.

Inflow Inflow

To PumpSuction

To PumpSuction

Baffle

FIGURE 5

Page 13: Centrifugal Pumps Handbook

The Pump Handbook Series 13

inimum flow can bedetermined by examin-ing each of the factorsthat affect it. There are

five elements that can be quanti-fied and evaluated:

1. Temperature rise (minimumthermal flow)

2. Minimum stable flow

3. Thrust capacity

4. NPSH requirements

5. Recirculation

The highest flow calculatedusing these parameters is consid-ered the minimum flow.

TEMPERATURE RISETemperature rise comes from

energy imparted to the liquidthrough hydraulic and mechanicallosses within the pump. Theselosses are converted to heat,which can be assumed to beentirely absorbed by the liquidpumped. Based on this assump-tion, temperature rise ∆T in °F isexpressed as:

H 1∆T = ————— x ——————

778 x Cp η – 1

where

H = total head in feet

Cp = specific heat of the liquid,Btu/lb x °F

η = pump efficiency in decimalform

778 ft–lbs = energy to raise thetemperature of one pound ofwater 1°F

To calculate this, the specificheat and allowable temperaturerise must be known.

The specific heat for water is1.0, and other specific heats canbe as low as 0.5. The specificheats for a number of liquids canbe found in many chemical and

mechanical handbooks.What is the maximum allowable

temperature rise? Pump manufactur-ers usually limit it to 15°F. However,this can be disastrous in certain situa-tions. A comparison of the vapor pres-sure to the lowest expected suctionpressure plus NPSH required (NPSHR)by the pump must be made. The tem-perature where the vapor pressureequals the suction pressure plus theNPSHR is the maximum allowable

temperature. The differencebetween the allowable temperatureand the temperature at the pumpinlet is the maximum allowabletemperature rise. Knowing ∆T andCp, the minimum flow can bedetermined by finding the corre-sponding head and efficiency.

When calculating the maxi-mum allowable temperature rise,look at the pump geometry. Forinstance, examine the vertical can

Elements of Minimum FlowBY TERRY M. WOLD

M

A high-pressure vertical pump. Asterisks (*) denote where low-temperature fluid is exposed to higher temperatures. Flashing andvaporization can occur here. Temperature increases as fluid trav-els from A towards B.

SUCTION

Low PressureLower

Temperature

DISCHARGE

High PressureHigherTemperature

CENTRIFUGAL PUMPS

HANDBOOK

FIGURE 1

Page 14: Centrifugal Pumps Handbook

14 The Pump Handbook Series

pump in Figure 1. Although pressureincreases as the fluid is pumpedupward through the stages, considerthe pump inlet. The fluid at the inlet(low pressure, low temperature) isexposed to the temperature of thefluid in the discharge riser in thehead (higher pressure, higher tem-perature). This means that the vaporpressure of the fluid at the pumpinlet must be high enough to accom-modate the total temperature risethrough all the stages. If this condi-tion is discovered during the pumpdesign phase, a thermal barrier canbe designed to reduce the tempera-ture that the inlet fluid is exposed to.

Some books, such as the PumpHandbook (Ref. 5), contain a typicalchart based on water (Cp = 1.0) thatcan be used with the manufacturer’sperformance curve to determinetemperature rise. If the maximumallowable temperature rise exceedsthe previously determined allowabletemperature rise, a heat shield canbe designed and fitted to the pumpduring the design stage. This require-ment must be recognized during thedesign stage, because once the pumpis built, options for retrofitting thepump with a heat shield are greatlyreduced.

MINIMUM STABLE FLOWMinimum stable flow can be

defined as the flow corresponding tothe head that equals shutoff head. Inother words, outside the ”droop“ inthe head capacity curve. In general,pumps with a specific speed lessthan 1,000 that are designed for opti-mum efficiency have a droopingcurve. Getting rid of this ”hump“requires an impeller redesign; how-ever, note that there will be a loss ofefficiency and an increase in NPSHR.

What’s wrong with a droopinghead/capacity curve? A droopingcurve has corresponding heads fortwo different flows. The pump reactsto the system requirements, andthere are two flows where the pumpcan meet the system requirements.As a result, it ”hunts“ or ”shuttles“between these two flows. This candamage the pump and other equip-ment, but it will happen only undercertain circumstances:

1. The liquid pumpedmust be uninhibitedat both the suctionand discharge ves-sels.

2. One element in thesystem must be ableto store and returnenergy, i.e., a watercolumn or trappedgas.

3. Something mustupset the system tomake it start hunt-ing, i.e., startinganother pump inparallel or throttlinga valve.

Note: All of thesemust be present at thesame time to cause thepump to hunt.

Minimum flowbased on the shape ofthe performance curveis not so much a func-tion of the pump as it isa function of the systemwhere the pump isplaced. A pump in a sys-tem where the abovecriteria are presentshould not have a droop-ing curve in the zone ofoperation.

Because pumps witha drooping head/capacitycurve have higher effi-ciency and a lower operating cost, itwould seem prudent to investigate theinstallation of a minimum flow bypass.

THRUST LOADINGAxial thrust in a vertical turbine

pump increases rapidly as flows arereduced and head increased. Based onthe limitations of the driver bearings,flow must be maintained at a valuewhere thrust developed by the pumpdoes not impair bearing life. Find outwhat your bearing life is and ask thepump manufacturer to give specificthrust values based on actual tests.

If a problem exists that cannot behandled by the driver bearings, con-tact the pump manufacturer. Thereare many designs available today forvertical pumps (both single and mul-

tistage) with integral bearings. Thesebearings can be sized to handle thethrust. Thrust can be balanced by theuse of balanced and unbalancedstages or adding a balance drum, ifnecessary. These techniques forthrust balancing are used when highthrust motors are not available. It isworth noting that balanced stagesincorporate wear rings and balanceholes to achieve lower thrust; there-fore, a slight reduction in pump effi-ciency can be expected, and energycosts become a factor.

NPSH REQUIREMENTSHow many pumps have been

oversized because of NPSH available(NPSHA)? It seems the easiest solu-tion to an NPSH problem is to go tothe next size pump with a larger suc-

FIGURE 2

Recirculation zones are always on the pres-sure side of the vane. A shows dischargerecirculation (the front shroud has been leftout for clarity). B shows inlet recirculation.

Page 15: Centrifugal Pumps Handbook

The Pump Handbook Series 15

Recirculation is caused by over-sized flow channels that allow liquidto turn around or reverse flow whilepumping is going on (Figure 2 showsrecirculation zones). This reversalcauses a vortex that attaches itself tothe pressure side of the vane. If thereis enough energy available and thevelocities are high enough, damagewill occur. Suction recirculation isreduced by lowering the peripheralvelocity, which in turn increasesNPSH. To avoid this it is better to rec-ognize the problem in the designstage and opt for a lower-speedpump, two smaller pumps, or anincrease in NPSHA.

Discharge recirculation iscaused by flow reversal and highvelocities producing damaging vor-tices on the pressure side of thevane at the outlet (Figure 2). Thesolution to this problem lies in the

tion, thereby reducing the inlet loss-es. A couple of factors become entan-gled when this is done. A largerpump means operating back on thepump curve. Minimum flow must beconsidered. Is the curve stable? Whatabout temperature rise? If there isalready an NPSH problem, an extrafew degrees of temperature rise willnot help the situation. The thrust andeye diameter will increase, possiblycausing damaging recirculation.When trying to solve an NPSH prob-lem, don’t take the easiest way out.Look at other options that may solvea long-term problem and reduce oper-ating costs.

RECIRCULATION

Every pump has a point whererecirculation begins. But if this is thecase, why don’t more pumps haveproblems?

10 15 20 25 30 35 40

impeller design. The problem is theresult of a mismatched case andimpeller, too little vane overlap inthe impeller design, or trimming theimpeller below the minimum diame-ter for which it was designed.

Recirculation is one of the mostdifficult problems to understand anddocument. Many studies on thetopic have been done over the years.Mr. Fraser’s paper (Ref. 1) is one ofthe most useful tools for determin-ing where recirculation begins. In ithe describes how to calculate theinception of recirculation based onspecific design characteristics of theimpeller and he includes charts thatcan be used with a minimumamount of information. An exampleof Fraser calculations, which showthe requirements to calculate theinception of suction and dischargerecirculation, is shown in Figure 3.

RECIRCULATION CALCULATIONS

Figure 3 indicates the user-defined variables and charts requiredto make the Fraser calculations forminimum flow. Information to do thedetailed calculations include:

Q = capacity at the best efficiency point

H = head at the best efficiencypoint

NPSHR = net positive suction headrequired at the pump suction

N = pump speedNS = pump specific speedNSS = suction specific speedZ = number of impeller vanesh1 = hub diameter (h1 = 0 for sin-

gle suction pumps)D1 = impeller eye diameterD2 = impeller outside diameterB1 = impeller inlet widthB2 = impeller outlet widthR1 = impeller inlet radiusR2 = impeller outlet radiusF1 = impeller inlet areaF2 = impeller outlet areaβ1 = impeller inlet angleβ2 = impeller outlet angle

The above information isobtained from the pump manufactur-er curves or impeller design files. Theimpeller design values are usuallyconsidered proprietary information.

KVe and KCm2 can be determinedfrom the charts in Figure 3.

FIGURE 3

Incipient recirculation. Minimum flow is approximately 50% ofincipient flow, while minimum intermittent flow is approximately25% of incipient flow. See text under “Recirculation Calculations”for details

Cm2U2

Discharge Angle β2 Inlet Angle β1

VeU1

.14

.12

.10

.08

.06

.04

10 15 20 25 30.02

.10

.12

.14

.16

.18

.20

.22

.24

.26

.28

.30

.32

R1

R2

D2

D1B1

B2

h1

.08

Page 16: Centrifugal Pumps Handbook

16 The Pump Handbook Series

With all of the above informa-tion at hand, suction recirculationand the two modes of dischargerecirculation can be determined.

As previously mentioned,Fraser has some empirical chartsat the end of his paper that can beused to estimate the minimumflow for recirculation. A few ofthe design factors of the impellerare still required. It is best to dis-cuss recirculation with yourpump manufacturer before pur-chasing a pump, in order toreduce the possibility of problemswith your pump and system afterinstallation and start-up.

SUMMARYMinimum flow can be accu-

rately determined if the elementsdescribed above are reviewed bythe user and the manufacturer.Neither has all the information todetermine a minimum flow that

is economical, efficient, and insuresa trouble-free pump life. It takes acoordinated effort by the user andthe manufacturer to come up withan optimum system for pump selec-tion, design, and installation.

REFERENCES1. F.H. Fraser. Recirculation in cen-

trifugal pumps. Presented at theASME Winter Annual Meeting(1981).

2. A.R. Budris. Sorting out flow recir-culation problems. Machine Design(1989).

3. J.J. Paugh. Head-vs-capacitycharacteristics of centrifugalpumps. Chemical Engineering(1984).

4. I. Taylor. NPSH still pump appli-cation problem. The Oil and GasJournal (1978).

5. I.J. Karassik. Pump Handbook.McGraw-Hill (1986). ■

Terry Wold has been the engi-neering manager for Afton Pumpsfor the last four years. He has beeninvolved in pump design for 25years. Mr. Wold graduated fromLamar Tech in 1968 with a bache-lor’s degree in mechanical engineer-ing and is currently a registeredengineer in the State of Texas.

Thanks to P.J. Patel for hiscomments and assistance in prepar-ing the graphics.

Page 17: Centrifugal Pumps Handbook

The Pump Handbook Series 17

ne of the greatest sourcesof power waste is the prac-tice of oversizing a pumpby selecting design condi-

tions with excessive margins inboth capacity and total head. It isstrange on occasion to encounter agreat deal of attention being paidto a one-point difference in effi-ciency between two pumps whileat the same time potential powersavings are ignored through anoverly conservative attitude inselecting the required conditionsof service.

POWER CONSUMPTION

After all, we are not primarilyinterested in efficiency; we aremore interested in power con-sumption. Pumps are designed toconvert mechanical energy from adriver into energy within a liquid.This energy within the liquid isneeded to overcome friction loss-es, static pressure differences andelevation differences at the desiredflow rate. Efficiency is nothing butthe ratio between the hydraulicenergy utilized by the process andthe energy input to the pump dri-ver. And without changing theratio itself, if we find that we areassigning more energy to theprocess than is really necessary,we can reduce this to correspondto the true requirement and there-fore reduce the power consump-tion of the pump.

It is true that some capacitymargin should always be includ-ed, mainly to reduce the wear ofinternal clearances which will,with time, reduce the effectivepump capacity. How much mar-gin to provide is a fairly complexquestion because the wear thatwill take place varies with thetype of pump in question, the liq-uid handled, the severity of theservice and a number of othervariables.

A centrifugal pump operatingin a given system will deliver acapacity corresponding to the

Effects of OversizingBY: IGOR J. KARASSIK

O

Pump H-Q curve superimposed on system-headcurve

intersection of itshead-capacity curvewith the system-head curve, as longas the availableNPSH is equal to orexceeds the requiredNPSH (Figure 1).To change this op-erating point in anexisting installationrequires changingeither the head-capacity curve orthe system-headcurve, or both. Thefirst can be accom-plished by varyingthe speed of thepump (Figure 2), orits impeller dia-meter while thesecond requiresaltering the frictionlosses by throttlinga valve in the pumpdischarge (Figure3). In the majorityof pump installa-tions, the driver is a constant speedmotor, and chang-ing the system-headcurve is used tochange the pumpcapacity. Thus, ifwe have providedtoo much excessmargin in the selec-tion of the pumphead-capacity curve,the pump will haveto operate with con-siderable throttlingto limit its deliveryto the desired value.

If, on the otherhand, we permitthe pump to oper-ate unthrottled,which is more like-ly, the flow into thesystem will increaseuntil that capacityis reached where

FIGURE 1

FIGURE 2

Varying pump capacity by varying speed

Varying pump capacity by throttling

FIGURE 3

CENTRIFUGAL PUMPS

HANDBOOK

H – Q CurveSystem-

Head Curve

CapacityH

ead

Head-Capacity at Full Speed (N1)

Head-Capacity at Full Speed (N2)Head-Capacity at Full Speed (N3) H3

H2

H1

System-Head Curve

FrictionLossesStaticPressur eor Head

}Hea

d

Capacity Q3 Q2 Q1max

Head-Capacity at Constant Speed H3

H2 H1

System-Head Curve

FrictionLossesStaticPressur eor Head

}

Capacity Q3 Q2 Q1max

SystemHead Curveby Throttling Valve

Hea

d

Page 18: Centrifugal Pumps Handbook

18 The Pump Handbook Series

the system-head and head-capaci-ty curves intersect.

EXAMPLE

Let’s use a concrete example,for which the maximum requiredcapacity is 2700 gpm, the statichead is 115 ft and the total frictionlosses, assuming 15-year old pipe,are 60 ft. The total head requiredat 2700 gpm is therefore 175 ft.We can now construct a system-head curve, which is shown oncurve A, Figure 4. If we add amargin of about 10% to therequired capacity and then, as isfrequently done, we add some

If we operate it throttled at therequired capacity of 2700 gpm,operating at the intersection of itshead-capacity curve and curve B,the pump will require 165 bhp.

The pump has been selectedwith too much margin. We cansafely select a pump with a small-er impeller diameter, say 14 in.,with a head-capacity curve asshown on Figure 4. It will inter-sect curve A at 2820 gpm, givingus about 4% margin in capacity,which is sufficient. To limit theflow to 2700 gpm, we will stillhave to throttle the pump slightlyand our system head curve willbecome curve C. The power con-sumption at 2700 gpm will now beonly 145 bhp instead of the 165 bhprequired with our first overly con-servative selection. This is a veryrespectable 12% saving in powerconsumption. Furthermore, we no longer need a 200 hp motor. A150 hp motor will do quite well.The saving in capital expenditureis another bonus resulting fromcorrect sizing.

Our savings may actually beeven greater. In many cases, thepump may be operated unthrot-tled, the capacity being permittedto run out to the intersection of thehead-capacity curve and curve A.If this were the case, a pump witha 14-3/4 in. impeller would operateat approximately 3150 gpm andtake 177 bhp. If a 14 in. impellerwere to be used, the pump wouldoperate at 2820 gpm and take 148 bhp. We could be saving morethan 15% in power consumption.Tables 1 and 2 tabulate these savings.

And our real margin of safetyis actually even greater than I haveindicated. Remember that the fric-tion losses we used to construct thesystem-head curve A were basedon losses through 15-year old piping. The losses through newpiping are only 0.613 times thelosses we have assumed. The sys-tem-head curve for new piping isthat indicated as curve D in Figure4. If the pump we had originallyselected (with a 14-3/4 in. impeller)were to operate unthrottled, itwould run at 3600 gpm and take

margin to the total head above thesystem-head curve at this rated flow,we end up by selecting a pump for3000 gpm and 200 ft. total head. Theperformance of such a pump, with a14-3/4 in. impeller, is superimposed onthe system-head curve A in Figure 4.

The pump develops excess headat the maximum required capacity of2700 gpm, and if we wish to operateat that capacity, this excess head willhave to be throttled. Curve “B” is thesystem-head curve that will have tobe created by throttling.

If we operate at 3000 gpm, thepump will take 175 bhp, and we willhave to drive it with a 200 hp motor.

Effect of oversizing a pump

B

A

D

C

H-Q 1800 R.P.M.

143/4"Impeller

η−Q

143 / 4"Im

pelle

r

η−Q143 /4"I

mpeller

14"Impeller

η−Q

14"Im

pelle

r

η−Q14"Impeller

Static Head

H-Q 1800 R.P.M.

0 1000 2000 3000 4000Capacity in G.P.M.

240

220

200

180

160

140

200

180

160

140

120

100

80

60

90

80

70

60

50

40

30

20

10

B.H

.P.

% E

ffici

ency

Fee

t Tot

al H

ead

FIGURE 4

Page 19: Centrifugal Pumps Handbook

The Pump Handbook Series 19

C l e a r l y ,important energysavings can beachieved if, at thetime of the selec-tion of the condi-tions of service,r e a s o n a b l erestraints are exer-cised to avoidi n c o r p o r a t i n gexcessive safetymargins into therated conditions ofservice.

EXISTINGINSTALLATIONS

But what ofexisting installationsin which the pumpor pumps haveexcessive margins?Is it too late toachieve these sav-ings? Far from it! Asa matter of fact, it ispossible to establishmore accurately thetrue system-head

curve by running a performance testonce the pump has been installed andoperated. A reasonable margin canthen be selected and several choicesbecome available to the user:

1. The existing impeller can be cutdown to meet the more realisticconditions of service.

2. A replacement impeller with thenecessary reduced diameter canbe ordered from the pump man-

ufacturer. The originalimpeller is then stored for fu-ture use if friction losses areultimately increased with timeor if greater capacities areever required.

3. In certain cases, there may betwo separate impeller designsavailable for the same pump,one of which is of narrowerwidth than the one originallyfurnished. A replacement nar-row impeller can then beordered from the manufactur-er. Such a narrower impellerwill have its best efficiency ata lower capacity than the nor-mal width impeller. It may ormay not need to be of a small-er diameter than the originalimpeller, depending on thedegree to which excessivemargin had originally beenprovided. Again, the originalimpeller is put away for possi-ble future use. ■

Igor J. Karassik is SeniorConsulting Engineer for Ingersoll-Dresser Pump Company. He hasbeen involved with the pump industryfor more than 60 years. Mr.Karassik is Contributing Editor -Centrifugal Pumps for Pumps andSystems Magazine.

187.5 bhp. A pump with only a 14 in. impeller would intersect thesystem-head curve D at 3230 gpmand take 156.6 bhp, with a savingof 16.5%. As a matter of fact, wecould even use a 13-3/4 in. impel-ler. The head-capacity curve wouldintersect curve D at 3100 gpm, andthe pump would take 147 bhp.Now, the savings over using a 14-3/4 in. impeller becomes 21.6%(See Table 3).

Throttled to 2700 GPMImpeller 143/4" 14"Curve “B” “C”BHP 165 145Savings 20 hp or 12.1%

TABLE 1. COMPARISON OF PUMPS WITH 143/4 IN. AND 14IN. IMPELLERS, WITH THE SYSTEM THROTTLED

Unthrottled, on Curve “A”Impeller 143/4" 14"GPM 3150 2820BHP 177 148Savings 29 hp or 16.4%

TABLE 2. COMPARISON OF PUMPS WITH THE SYSTEM UNTHROTTLED

Impeller 143/4" 14" 133/4"GPM 3600 3230 3100BHP 187.5 156.5 147Savings 31 hp 40.5 hp

16.5% 21.6%

TABLE 3. EFFECT OF DIFFERENT SIZE IMPELLERS IN SYSTEM WITH NEW PIPE AND RESULTING SAVINGS NEW PIPE (UNTHROTTLED OPERATION, CURVE “D”)

Page 20: Centrifugal Pumps Handbook

20 The Pump Handbook Series

hen sizing a pump for anew application or eval-uating the performanceof an existing pump, it is

often necessary to account for theeffect of the pumped fluid’s vis-cosity. We are all aware that thehead-capacity curves presented inpump vendor catalogs are pre-pared using water as the pumpedfluid. These curves are adequatefor use when the actual fluid thatwe are interested in pumping hasa viscosity that is less than orequal to that of water. However,in some cases—certain crude oils,for example—this is not the case.

Heavy crude oils can haveviscosities high enough to increasethe friction drag on a pump’simpellers significantly. The addi-tional horsepower required toovercome this drag reduces thepump’s efficiency. There are sev-eral analytical and empiricalapproaches available to estimatethe magnitude of this effect. Someof these are discussed below.

Before beginning the discus-sion, however, it is vital to empha-size the importance of having anaccurate viscosity number onwhich to base our estimates. Theviscosity of most liquids is strong-ly influenced by temperature. Thisrelationship is most often shownby plotting two points on a semi-logarithmic grid and connectingthem with a straight line. The rela-tionship is of the form:

µ = AeB/T

where

µ = the absolute viscosity of thefluid

A and B = constants

T = the absolute temperature of thefluid

Plotting this relationshiprequires knowledge of two datapoints, and using them effective-ly requires some judgement as to

the normal operating temperatureas well as the minimum tempera-ture that might be expected duringother off-design conditions such asstart-up.

The effect of pressure on theviscosity of most fluids is small.For mineral oils, for example, anincrease of pressure of 33 bars(≈480 psi ) is equivalent to a tem-

Fluid Viscosity Effects on Centrifugal PumpsBY: GUNNAR HOLE

W FIGURE 1

Reproduced from the Hydraulic Institute Standards (Figure 71)

CENTRIFUGAL PUMPS

HANDBOOK

Page 21: Centrifugal Pumps Handbook

The Pump Handbook Series 21

perature drop of 1°C.The following definitions are

used when discussing fluids andviscosity. There are five basictypes of liquid that can be differ-entiated on the basis of their vis-cous behavior; they are:

NON-NEWTONIANThese are fluids where the

shear rate-shear stress relationshipis nonlinear. They can be dividedinto four categories:• Bingham-plastic fluids are

those in which there is noflow until a threshold shearstress is reached. Beyond thispoint, viscosity decreases withincreasing shear rate. Mostslurries have this property, asdoes America’s favorite veg-etable, catsup.

• Dilatant fluids are those ofwhich viscosity increaseswith increasing shear rate.Examples are candy mixtures,clay slurries, and quicksand.

• Pseudo-plastic fluids are simi-lar to Bingham-plastic fluids,except there is no definiteyield stress. Many emulsionsfall into this category.

• Thixotropic fluids are those ofwhich viscosity decreases to aminimum level as their shearrate increases. Their viscosityat any particular shear ratemay vary, depending on theprevious condition of the fluid.Examples are asphalt, paint,molasses, and drilling mud.

There are two other termswith which you should be familiar:

• Dynamic or absolute viscosityis usually measured in termsof centipoise and has the unitsof force time/length2.

• Kinematic viscosity is usuallymeasured in terms of centis-tokes or ssu (Saybolt SecondsUniversal). It is related toabsolute viscosity as follows:

kinematic viscosity = absolute viscosity/mass density

The normal practice is for thisterm to have the units of length2/time. Note:

1 cSt = cP x sp gr

1 cSt = 0.22 x ssu – (180/ssu)

1 cP = 1.45E-7 lbf – s/in2

1 Reyn = 1 lbf – s/in2

NEWTONIANThese are fluids where viscosity is

constant and independent of shearrate, and where the shear rate is linear-ly proportional to the shear stress.Examples are water and oil.

FIGURE 2

Reproduced from the Hydraulic Institute Standards (Figure 72 )

Page 22: Centrifugal Pumps Handbook

22 The Pump Handbook Series

The process of determiningthe effect of a fluid’s viscosity onan operating pump has been stud-ied for a number of years. In thebook Centrifugal and Axial FlowPumps, A.J. Stepanoff lists thelosses that affect the performanceof pumps as being of the follow-ing types:

• mechanical losses

• impeller losses

• leakage losses

• disk friction losses

Of all external mechanicallosses, disk friction is by far themost important, according toStepanoff. This is particular-ly true for pumps designed withlow specific speeds. Stepanoffgives a brief discussion of thephysics of a rotating impeller andemerges with a simple equationthat summarizes the drag forceacting upon it:

(hp)d = Kn3D5

where

K = a real constant

n = the pump operating speed

D = the impeller diameter

The explana-tion further de-scribes the motionof fluid in theimmediate neigh-borhood of thespinning impeller.There Stepanoffmentions the exper-imental results ofothers demonstrat-ing that, by reduc-ing the clearancebetween the sta-tionary casing andthe impeller, the re-quired power canbe reduced. Healso writes about

the details of some investigationsthat demonstrate the beneficialeffect of good surface finishes onboth the stationary and rotating sur-faces. Included is a chart preparedby Pfleiderer, based on work byZumbusch and Schultz-Grunow,that gives friction coefficients forcalculating disk friction losses. Thechart is used in conjunction withthe following equation:

(hp)d = KD2γu3

where

K = a constant based on the Reynoldsnumber

D = impeller diameter

γ= fluid density

u = impeller tip speed

Like most of Stepanoff’s writing,this presentation contains great depthwith considerable rigor. It makesinteresting reading if you are willingto put forth the time. Those of us

who need a quick answer to a par-ticular problem may need to lookelsewhere for help.

In the book, CentrifugalPumps, V. Lobanoff and R. Rossdiscuss the effect of viscous fluidson the performance of centrifugalpumps. They make the point thatbecause the internal flow pas-sages in small pumps are propor-tionally larger than those in largerpumps, the smaller pumps willalways be more sensitive to theeffects of viscous fluids. Theyalso introduce a diagram from thepaper “Engineering and SystemDesign Considerations for PumpSystems and Viscous Service,” byC.E. Petersen, presented atPacific Energy Association,October 15, 1982. In this dia-gram, it is recommended that themaximum fluid viscosity a pumpshould be allowed to handle belimited by the pump’s dischargenozzle size. The relationship isapproximately:

viscositymax = 300(Doutlet nozzle –1)

where

viscosity is given in terms of ssu

D is measured in inches

With respect to the predictionof the effects of viscous liquids onthe performance of centrifugalpumps, Lobanoff and Ross directthe reader to the clearly definedmethodology of the HydraulicInstitute Standards. This techniqueis based on the use of two nomo-grams on pages 112 and 113 of the14th edition (Figures 71 and 72).They are reproduced here asFigures 1 and 2. They are intended

WaterCurve-BasedPerformance % of BEP Capacity

60% 80% 100% 120%Capacity, gpm 450 600 750 900Differential Head, ft. 120 115 100 100Efficiency 0.70 0.75 0.81 0.75Horsepower 18 21 21 27Viscous (1,000 ssu)PerformanceCapacity, gpm 423 564 705 846Differential Head, ft. 115 108 92 89Efficiency 0.45 0.48 0.52 0.48Horsepower 25 29 28 36

TABLE 1. WATER-BASED AND VISCOUS PERFORMANCE

Note: Pumped fluid specific gravity = 0.9

CorrectionFactor

Dx1 Dx2 Dx3 Dx4 Dx5 Dx6

Cη 1.0522 -3.5120E-02 -9.0394E-04 2.2218E-04 -1.1986E-05 1.9895E-07CQ 0.9873 9.0190E-03 -1.6233E-03 7.7233E-05 -2.0528E-06 2.1009E-08

CH0.6 1.0103 -4.6061E-03 2.4091E-04 -1.6912E-05 3.2459E-07 -1.6611E-09CH0.8 1.0167 -8.3641E-03 5.1288E-04 -2.9941E-05 6.1644E-07 -4.0487E-09CH1.0 1.0045 -2.6640E-03 -6.8292E-04 4.9706E-05 -1.6522E-06 1.9172E-08CH1.2 1.0175 -7.8654E-03 -5.6018E-04 5.4967E-05 -1.9035E-06 2.1615E-08

TABLE 2. POLYNOMIAL COEFFICIENTS

Page 23: Centrifugal Pumps Handbook

for use on pumps with BEPsbelow and above 100 gpm, respec-tively, which permits the user toestimate the reduction of head,capacity, and efficiency that a vis-cous fluid will produce on a pumpcurve originally generated withwater. A variation on this tech-nique is described below.

The following example istaken from pages 114-116 of theHydraulic Institute Standards sec-tion on centrifugal pump applica-tions. There, the use of Figure 72,“Performance Correction ChartFor Viscous Liquids,” is discussed.Table 1 was calculated using poly-nomial equations developed toreplace the nomogram presentedin Figure 72. The results of the cal-culation are within rounding errorof those presented in the standard.And the approach has the addi-tional benefit of being more conve-nient to use, once it has been setup as a spreadsheet.

In the course of curve-fittingFigure 72, it was convenient todefine a term known as pseudoca-pacity:

pseudocapacity =1.95(V)0.5[0.04739(H)0.25746(Q)0.5]-0.5

where

V = fluid viscosity in centistokes

H = head rise per stage at BEP, mea-sured in feet

Q = capacity at BEP in gpm

Pseudocapacity is used with thefollowing polynomial coefficients todetermine viscosity correction termsthat are very close to those given byFigure 72 in the Hydraulic InstituteStandards. These polynomials havebeen checked throughout the entirerange of Figure 72, and appear to giveanswers within 1.0% of those foundusing the figure.

The polynomial used is of theform:

Cx = Dx1 + Dx2P + Dx3P2 + Dx4P3 +Dx5P4 + Dx6P5

where

Cx is the correction factor that must beapplied to the term in question

Dxn are the polynomial coefficients listedin Table 2

P is the pseudocapacity term definedabove

For comparison, the correctionfactors for the example above (tabu-lated in Table 7 of the HydraulicInstitute Standards) and those calculat-ed using the polynomial expressionsabove are listed in Table 3.

The problem of selecting a pumpfor use in a viscous service is relative-ly simple once the correction coeffi-cients have been calculated. If, forexample, we had been looking for apump that could deliver 100 feet of

head at a capacity of 750 gpm, wewould proceed as follows:

Hwater = Hviscous service/CH1.0

Qwater = Qviscous service/CQ

The next step would be tofind a pump having the requiredperformance on water. Afterdetermining the efficiency of thepump on water, we would correctit for the viscous case as shownabove:

ηviscous service = ηwater x Cη

The horsepower required bythe pump at this point would becalculated as follows:

hpviscous service =

(Qviscous service x Hviscous service x sp gr)

(3,960 x ηviscous service)

As with water service, thehorsepower requirements at off-design conditions should alwaysbe checked. ■

Gunnar Hole is a principal inTrident Engineering, Inc. inHouston, TX. He has been involvedin the selection, installation, andtroubleshooting of rotating equip-ment for the past 15 years. Mr.Hole is a graduate of the Universityof Wisconsin at Madison and is aRegistered Professional Engineer inTexas.

The Pump Handbook Series 23

Cη CQ CH0.6 CH0.8 CH1.0 CH1.2

Per Table 7 of HI Standards 0.635 0.95 0.96 0.94 0.92 0.89Per Polynomial Expressions 0.639 0.939 0.958 0.939 0.916 0.887

TABLE 3. CORRECTION FACTOR COMPARISON

Page 24: Centrifugal Pumps Handbook

24 The Pump Handbook Series

Presented below is a descriptionof the problem, definitions of some ofthe more important terms used, andreferences that can be consulted for amore thorough review. A table alsocompares three of the most commonbalancing criteria used in the pumpindustry.

Perhaps the least controversialcomment that can be made to anexperienced equipment specialist isthat “accurate rotor balancing is criti-cal to reliable operation.” I could addsome spice to the conversation by giv-ing my opinion on how good is goodenough, but I would rather address

the standards used in the pumpindustry and show how they takedifferent approaches to resolve theproblem of balancing rotors.

I use the term rotor repeated-ly in this discussion. For the pur-pose of this article, I includepartially and fully assembledpump shaft/sleeve/impeller as-semblies as well as individualpump components installed onbalancing machine arbors in thisdefinition.

The three major criteria usedwill be referred to as theUnbalanced Force Method (UFM),the Specified Eccentricity Method(SEM), and the Specified CircularVelocity Method (SCVM).

In the UFM the allowableunbalance permitted in a rotor isthe amount that will result in adynamic force on the rotor systemequal to some percentage of therotor’s static weight. This allow-able unbalance is therefore relatedto the operating speed of the rotor.An example of this method can befound in API Standard 610 6thEdition, where the unbalanceforce contributed to a rotor systemby a rotating unbalance is limitedto 10% of the rotor’s static weight.

The SEM attempts to specifybalance quality by limiting thedistance by which the center ofmass of the rotor can be offsetfrom the center of rotation of therotor. This method is used inAGMA Standard 515.02, which is

Unbalanced Specified Specified Force Eccentricity Circular

Method Method VelocityAs per API 610 As per Method

6th Edition AGMA 510.02 As per API 6107th Edition

Residual Unbalance(RUB), in.–oz 56347 Wj 16 εWj 4 Wj

where: N2 NWj = rotor weight per balance plane, Ibf N = rpmε= eccentricity, in.Eccentricity (ε) or Specific Unbalancein.–oz/lbm 56347 16 ε 4

N2 N

in.–lbm/lbm 3522 ε 0.25N2 N

where RUB = εWj see Table 2Unbalance Force (UBF),lbf where:UBF = εMω2 0.10 Wj εWjN2 WjNand M = Wj/386 lbf–s2/in. 35200 140800ω = 2 π N/60 rad/sCircular Velocity (CV),in./s 368 εN 0.26

N 9.54

mm/s 9347 2.66 εN 0.665N

where CV = εω ISO Standard 1940 G – 9347 G – 2.66 εN G – 0.665Balance Grade N

he subject of balancingrotors is one of the funda-mentals of rotating equip-ment engineering. A

number of balancing standardshave been developed over theyears to meet the requirements ofpump manufacturers and users,and the idea of balancing is simple.Unfortunately, the definitions andmathematics used in describingbalancing problems can be confus-ing. This article compares thesecriteria so the end user can useconsistent reasoning when makingbalancing decisions.

commonly referenced by flexiblecoupling vendors. It has theadvantage of being conceptuallysimple. For the gear manufactur-ers who developed this standard,it allowed the use of manufactur-ing process tolerances as balanc-ing tolerances. In Paragraph 3.2.7,API 610 7th Edition suggests thatcouplings meeting AGMA 515.02Class 8 should be used unless oth-erwise specified.

The SCVM is based on con-siderations of mechanical similari-ty. For geometrically similar rigidrotors running at equal peripheralspeeds, the stresses in the rotorand bearings are the same. Thismethod is described in ISOStandard 1940—Balance Quality ofRigid Rotors. It also forms thebasis of API Standard 610 7thEdition’s very stringent 4W/N bal-ancing requirement. Standardsbased on this methodology arebecoming more common.

In Table 1 the three balancingcriteria discussed above are com-pared with respect to their effect onthe various parameters involved inbalancing. The terms used in thetable are defined as follows:

RESIDUAL UNBALANCEThis is the amount of unbal-

ance present or allowed in therotor. It has the units of mass andlength. It is computed by takingthe product of the rotor mass (perbalance plane) times the distancefrom the rotor’s center of mass toits center of rotation. Note that 1in.–oz is equivalent to 72.1 cm–g.

ECCENTRICITYThis is the distance that the

center of mass of the rotor is dis-placed from the rotor’s center ofrotation. It has the unit of length.It can also be considered as a mea-sure of specific residual unbal-ance, having the units oflength–mass/mass. This term isthe basic criterion of SEM balanc-ing rules (see Table 2). Note that 1in. is equivalent to 25.4 mm.

Pump Balancing CriteriaBY GUNNAR HOLE

T

TABLE 1. BALANCING CRITERIA

CENTRIFUGAL PUMPS

HANDBOOK

Page 25: Centrifugal Pumps Handbook

The Pump Handbook Series 25

FLEXIBLE ROTORThe elastic deflection of flexible

rotors sets up additional centrifugalforces that add to the original unbal-ance forces. Such rotors can be bal-anced in two planes for a single speedonly. At any other speed they willbecome unbalanced. Balancing therotor to allow it to run over a range ofspeeds involves corrections in threeor more planes. This process is calledmulti-plane balancing.

One important point is that thepump/coupling/driver system mustbe considered as a whole when eval-uating balance quality. A simplepump rotor can be balanced to meetAPI 610 7th Edition’s 4W/N criteriain a modern balancing machine with-out too much trouble. An electricmotor rotor may be even easier tobalance due to its simple construc-tion. But the coupling connectingthem can be a completely differentmatter.

The coupling will likely havemore residual unbalance than eitherthe pump or the motor. And everytime you take the coupling apart andput it back together you take thechance of changing its balance condi-tion. As written, API 610 7th Editionallows a coupling to have a specificresidual unbalance nearly 60 timeshigher than for a 3,600 rpm pump.This can be a significant problem ifyou use a relatively heavy coupling.

These balancing methods are pri-marily intended for use on rigidrotors—those operating at speedsunder their first critical speed.Flexible rotors, which operate abovetheir first critical speed, are consider-ably more complicated to balance.The process of balancing flexiblerotors is discussed in ISO Standard5406–The Mechanical Balancing ofFlexible Rotors and ISO Standard5343–Criteria for Evaluating FlexibleRotor Unbalance.

The basic concepts of rigid andflexible rotor balancing are the same.The main difference is that with rigidrotor balancing we are only con-cerned with the rigid body modes ofvibration. With a flexible rotor, wehave to consider some of the highermodes of vibration as well. In thesecases the deflection of the rotor affectsthe mass distribution along its length.In general, each of the modes has tobe balanced independently.

Note: AGMA 515.02 refers to several BalanceQuality Classes. They are summarized as follows:

Equivalent ISO AGMA Balance Quality GradeClass ε, µ-in. 1,800 rpm 3,600 rpm

8 4,000 19.2 38.39 2,000 9.6 19.210 1,000 4.8 9.611 500 2.4 4.812 250 1.2 2.4

UNBALANCE FORCEThis is the force that is exert-

ed on a rotor system as a result ofthe non-symmetrical distributionof mass about the rotor’s center ofrotation. The units of this term areforce. This term is the basic criteri-on of UFM balancing rules. Notethat 1 lbf is equivalent to 4.45Newton.

CIRCULAR VELOCITYThis is the velocity at which

the center of mass of the rotorrotates around the center of rota-tion. You can think of it as a tan-gential velocity term. It has theunits of length per unit time. Itforms the basis for balancing rulesbased on the ISO Standard 1940series. In fact, the BalancingGrades outlined in ISO 1940 arereferenced by their allowable circu-lar velocity in millimeters per sec-ond. The balance quality called forin API 610 7th Edition is betterthan the quality that ISO 1940 rec-ommends for tape recorder drivesand grinding machines. ISO 1940recommends G–6.3 and G–2.5 formost pump components, whereAPI 610 calls for the equivalent ofG–0.67. Note that 1 in./s is equiva-lent to 25.4 mm/s.

RIGID ROTORA rotor is considered rigid

when it can be balanced by mak-ing mass corrections in any twoarbitrarily selected balancingplanes. After these corrections aremade, the balance will not signifi-cantly change at any speed up tothe maximum operating speed.With the possible exception ofhome ceiling fans, I believe thattwo-plane balancing is the mini-mum required for rotating equip-ment components.

Appendix I of API 610 7thEdition briefly discusses some ofthe implications of operating arotor near a critical speed. Theguidelines given there recom-mend separation margins thatspecify how far away from a criti-cal speed you can operate a rotor.These margins depend on the sys-tem amplification factors (alsoknown as magnification factors),which are directly related to thedamping available for the mode orresonance in question. The netresult of these recommendationsis to limit the maximum operatingamplification factor to a maxi-mum of about 3.75. The amplifi-cation factor can be thought of asa multiplier applied to the masseccentricity, ε, to account for theeffect of system dynamics.Algebraically, the physics of thesituation can be represented asfollows:

x = X sin (ωt – Φ)

(ω/ωn)2

X = ε ————————————([1 – (ω/ωn)2]2 + (2ζω/ωn)2)0.5

2ζω/ωnΦ = tan–1 ————————

1 – (ω/ωn)2wherex is the displacement of a point on

the rotorX is the magnitude of the vibration

at that pointε is the mass eccentricityω is the operating speed or fre-

quency of the rotorΦ is the phase angle by which the

response lags the forceζ is the damping factor for the

mode of vibration under consid-eration

X/ε is the amplification factorω/ωn is the ratio of operating speed

to the critical speed under con-siderationA more detailed discussion on

the topic of damped unbalanceresponse (or whirling of shafts)can be found in any introductoryvibration textbook. ■

Gunnar Hole is a principal inTrident Engineering, Inc. in Houston,TX.

TABLE 2. BALANCE QUALITY CLASSES

Page 26: Centrifugal Pumps Handbook

26 The Pump Handbook Series

ntifriction bearings, whichcan utilize either balls orrollers, are used to transferradial and axial loads

between the rotating and station-ary pump and motor assembliesduring operation. Even under thebest of installation, maintenance,and operating conditions, bearingfailures can and will occur. Thepurpose of this article is to providea working-level discussion of bear-ings, the types of failures, andhow bearings should be installedand maintained for optimum lifeexpectancy.

Due to space limitations, wecannot address all the differentsizes and types of bearings avail-able, or all the constraints cur-rently utilized in design.However, because electric motorsare used more often to drive cen-trifugal pumps, our discussionwill be based on bearings typical-ly used in quality motors. Thesebearings usually include a singleradial bearing and a matched setof duplex angular contact bear-ings (DACBs). Together, thesebearings must:• allow the unit to operate satis-

factorily over long periods oftime with minimum frictionand maintenance

• maintain critical tolerancesbetween rotating and stationaryassemblies to prevent contactand wear

• transmit all variable radial andaxial loads in all operating con-ditions, which include reverserotation, startup, shutdown,maximum flow, and maximumdischarge pressure

Each bearing has a specificpurpose. The radial bearing,which is located at one end of themotor, only transfers radial loadssuch as minor unbalanced rotorloads—and the weight of the rotoritself in the case of horizontallyoriented components. The DACBs

must transfer radial loads at theother end of the motor, and theymust transfer all axial loads. Photo 1shows several typical radial bearings,and Photo 2 shows DACBs.

DIFFERENT BEARINGCONFIGURATIONS

Radial bearings may be providedwith either 0, 1, or 2 seals or shieldsthat are effectively used to prevententry of foreign material into thebearings. If the bearing is equippedwith one seal or shield, the installershould determine which end of themotor the seal or shield should face.Failure to install radial bearings prop-erly in the correct orientation mayresult in the blockage of grease orlubricant to the bearings during rou-tine maintenance.

The orientation of DACBs ismore complex, DACBs must beinstalled in one of four configura-tions, as determined by design:1. face-to-face2. back-to-back3. tandem: faces toward the pump4. tandem: faces away from the pump

The “face” of the DACB is thatside that has the narrow lip on the

outer race. The “back” of the bear-ing has the wider lip on the outerrace and usually has various sym-bols and designators on it. Photo 2shows two pairs of DACBs. Thepair on the left is positioned face-to-face while the pair on the rightis back-to-back. Note that the lipon the outer races of the first pairis narrower than on the secondpair. This distinguishing character-istic provides an easy identifica-tion of which side is the face orback. In tandem, the narrow lip ofone bearing is placed next to thewide lip on the other. In otherwords, all bearing faces pointeither toward the pump or awayfrom it.

To facilitate the installation ofDACBs, the bearing faces shouldbe marked with a black indeliblemarker showing where the bur-nished alignment marks (BAMs)are on the back. This is becausethe four BAMs, two on each bear-ing, must be aligned with theircounterparts, and not all BAMsare visible during installation. Forexample, when the first bearing isinstalled in a face-to-face configura-tion, the BAMs are on the back

Bearing BasicsBY RAY RHOE

A

Photo 1. Typical radial bearings

CENTRIFUGAL PUMPS

HANDBOOK

Page 27: Centrifugal Pumps Handbook

The Pump Handbook Series 27

side, hidden from the installer.Marking the face of each bearingallows the installer to see where theBAMs are, so that all four BAMsmay be aligned in the same relativeposition, such as 12 o’clock.

BEARING PRELOADUnder certain operating con-

ditions (hydraulic forces, gravity,and movement of the pump andmotor foundation such as on aseagoing vessel), the rotor may beloaded in either direction. If thisoccurs, the balls in a DACB withno preload could becomeunloaded. When this happens,the balls tend to slide against theraces (ball skid) rather than roll.This sliding could result in per-manent damage to the bearingsafter about five minutes.

To prevent ball skid, bearingmanufacturers provide bearingsthat have a predetermined clear-ance between either the inner orouter races. Face-to-face bearingshave this clearance between theouter races. When the bearings areclamped together at installation(the outer races are clampedtogether), the balls are pressedbetween the inner and outer races,causing the preload. Back-to-backbearings have the clearance on theinner races, which are usuallyclamped together with a bearinglocknut.

By increasing the clearancebetween races, the preload can beincreased from zero to a heavyload. This way, when conditions

cause the rotor loads to changedirection or be eliminated, the bear-ing balls will still be loaded and ballskid should not occur.

One disadvantage of using pre-loaded bearings is that bearing lifewill be reduced due to the increasedloading. Preloaded bearings shouldnot be used unless design conditionsrequire them.

If uncertain about the need forpreload, users should contact manu-facturers.

BEARING INSTALLATIONOnce the proper bearings have

been obtained and the correct orien-tation determined, installation is rela-tively simple.

The shaft and especially theshaft shoulder should be cleanedand any welding or grinding opera-tions secured. The bearings must beinstalled in a clean environment,and the shaft must be free of nicksand burrs that may interfere withinstallation.

1. RADIAL BEARINGSTo install radial bearings, they

should be heated in a portable ovento 180–200°F. Then, using cleangloves and remembering the correctorientation, quickly slide each bear-ing over the shaft and firmly onto theshaft shoulder. Do not drop or slapthem into position. Experience indi-cates that you have about 10 secondsafter removing the bearing from theoven before it cools and seizes theshaft. If it seizes the shaft out of loca-tion, remove it and scrap it. The bear-

ing cannot be hammered into posi-tion or removed and reusedbecause it will be destroyed inter-nally by these actions.

2. DACBsInstallation of DACBs follows

the some procedure, except thatadditional care must be taken toposition the bearings properly,line up the burnished alignmentmarks, and not erase the indeliblemarks added on each bearingface. After the first bearing hasbeen installed, rotate the rotor (ifnecessary) so the alignment markon the inner race is at 12 o’clock,then rotate the outer race so it toois at 12 o’clock. Before proceedingwith the second bearing, mentallywalk through the procedure.Remember which direction theface goes and that the burnishedalignment marks must be in thesame position as the first bearingmarks. Also remember you haveabout 10 seconds before the bear-ing seizes the shaft.

The purpose of aligning thefour burnished alignment marksis to minimize off-loading (fight)and radial runout loads that willoccur if the true centers of thebearings are not lined up. Minorimperfections will always occur,and they must be minimized.Failure to align the marks willresult in the bearings loadingeach other.

DACBs come only inmatched pairs—they must be usedtogether. To verify that a pair ismatched, check the serial numberon the bearing halves—theyshould be the same, or properlydesignated, such as using bearing“A” and bearing “B.”

NEW BEARING RUN-INAfter new bearings have been

installed, they should be run inwhile monitoring their tempera-ture, noise, and vibration. Run-inis often called the “heat run” or“bearing stabilization test.”

To perform this test, firstrotate the pump and driver byhand to check for rubbing or bind-ing. If none occurs, operate the

Photo 2. Two pairs of DACBs, with the pair on the left positionedface-to-face, the pair on the right back-to-back

Page 28: Centrifugal Pumps Handbook

28 The Pump Handbook Series

unit at the design rating point andrecord bearing temperaturesevery 15 min. Bearing tempera-tures should increase sharply andthen slowly decline to their nor-mal operating temperature, usual-ly 20–60°F above ambient.

During the heat run careshould be taken to ensure thatthe temperature does not exceedthe value specified by the manu-facturer. If it does, the unitshould be secured and allowed tocool to within 20°F of ambient,or for 2 hours. The unit may thenbe restarted and the test repeatedas necessary until the bearingtemperature peaks and begins todecline.

If, after repeated attempts,bearing temperatures do not showsigns of stabilization, too muchgrease may be present. The bear-ing should be inspected and cor-rective actions taken as necessary.

Now let’s cover some basicrules to follow when workingwith bearings:1. Never reuse a bearing that has

been removed using a gearpuller, even if it is new. Thebearing has been internallydestroyed in the removalprocess. See “True Brinelling”under “Bearing Failures.”

2. DACBs must be installed in thecorrect orientation. If not, theymay experience reverse loadingand fail. See “Reverse Loading”under “Bearing Failures.”

3. Bearings must be installed in aclean environment. Contami-nation is a leading cause of pre-mature failures.

4. Do not pack the bearing andbearing cups full of grease.Excessive grease will cause over-heating and ball skid. See“Excessive Lubrication” under“Bearing Failure.”

BEARING FAILURESFailure to follow these four

basic rules will result in prematurebearing failures. These and otherfailures will occur for the followingreasons:

True Brinelling (failure to fol-low Rule 1): This type of bearingfailure occurs when removing bear-ings with a gear puller. The forcerequired to remove a bearing froma shrink-fit application is greatenough that when it is transferredthrough the balls to the inner races,the balls are pressed into the innerand outer races forming permanentindentations.

Reverse Loading (failure tofollow Rule 2): Reverse loadingoccurs when DACBs are improperly

installed and the balls ride on theball ridge located on the outerrace. Evidence of reverse loadingappears as “equator” bandsaround the balls.

Contamination (failure tofollow Rule 3): Contamination ofbearings almost always occursduring installation, but can alsooccur when liquids or other con-stituents from the pump leak orare present in the surroundingenvironment. If contamination isfound in a new bearing beforeinstallation, the bearing should becarefully cleaned and repacked.Evidence of solid contamination inused bearings usually appears asvery small, flat dents in the racesand balls.

Excessive Lubrication (fail-ure to follow Rule 4): Too muchgrease in a bearing may cause theballs to “plow” their way throughthe grease, resulting in increasedfriction and heat. If the bearingsand bearing caps are packed fullof grease, ball skid could occur.When it does, the balls do not roll,but actually slide against the races.Experience shows that the bear-ings may be permanently damagedafter more than five minutes ofball skid. Finding packed bearingsand bearing caps is a good indica-tion that too much grease causedthe bearing to fail. Bearing manu-facturers usually recommend thatbearings have 25-50% of their freevolume filled with grease.

Excessive Heat: Failure toprovide adequate heat transferpaths, or operating the componentat excessive loads or speeds mayresult in high operating tempera-tures. Evidence of excessive tem-perature usually appears assilver/gold/brown/blue discol-oration of the metal parts.

False Brinelling: Falsebrinelling occurs when excessivevibrations cause wear and break-down of the grease film betweenthe balls and the races. This condi-tion may be accompanied by signsof corrosion. A good example ofhow false brinelling could occurwould be when a horizontallypositioned component is shipped

Photo 3. Radial bearing disassembly

Page 29: Centrifugal Pumps Handbook

The Pump Handbook Series 29

across the country and not cush-ioned from a rough road surface.The load of the rotor is passedthrough the bearing balls, whichwear away or indent the races.Evidence of false brinelling lookssimilar to true brinelling, but maybe accompanied by signs of corro-sion where the grease film has notbeen maintained. Correction sim-ply involves protecting the unitfrom excessive vibration and usingspecially formulated greaseswhere past experience demon-strates the need.

Fatigue Failure: Even whenall operating, installation, andmaintenance conditions are per-fect, bearings will still fail. In thiscase, the bearings have simplyreached the end of their usefullife, and any additional use resultsin metal being removed from theindividual components. Evidenceof fatigue failure appears as pits.

BEARING DISASSEMBLY FORINSPECTION

Now that we know what tolook for in failed bearings, let’s see

how we disassemble a bearingfor inspection. Before disassem-bling any bearing, however, turnit by hand and check it for roughperformance. Note its generalcondition, the grease (and quanti-ty thereof), and whether there isany contamination. If solid conta-mination is present, the particlesshould be collected using a cleanfilter bag as follows:1. Partially fill a clean bucket or

container with clean dieselfuel or kerosene.

2. Insert a clean filter bag into thekerosene container. Thisensures that no contaminationfrom the container or thekerosene gets into the filter

bag.3. Using a clean brush, wash the

grease and contamination outof the bearing. The grease will

dissolve and any contaminationwill be collected in the filter bagfor future evaluation.

RADIAL BEARING DISASSEMBLYAfter removing the grease and

any contamination, you should disas-semble radial bearings by removingany seals or shields, which are oftenheld in place by snap rings. Then, toremove a metal retainer, drillthrough the rivets and remove bothretainer halves. Then the bearingshould again be flushed (in a differ-ent location) to remove any metalshavings that may have fallenbetween the balls and races whendrilling out the rivets. If the bearingdoes not freely turn by hand, somemetal particles are still trappedbetween the balls and races.

Next, place the bearing on thefloor as shown in Photo 3 with theballs packed tightly together on thetop. Insert a rod or bar through theinner race and press down, hard ifnecessary. Note: If the balls are notpacked tightly together, disassemblywill not occur.

DACB DISASSEMBLYTo disassemble DACBs, sup-

port the face of the outer race andpress down against the inner race.The back of the bearing must beon top.

HANDLING, TRANSPORTATION,AND STORAGE

Common sense applies inhandling, storing, or transportingprecision bearings. They shouldnot be dropped or banged. Theyshould be transported by hand incushioned containers, or on theseat of vehicle—not in a bike rack.They should be stored in a cool,clean, dry environment.

Because nothing lasts forever,including bearing grease, bearingsshould not be stored for more thana few years. After this, the greasedegrades and the bearings maybecome corroded. At best, an oldbearing may have to be cleanedand repacked, using the correcttype and amount of grease.

MAINTENANCERoutine maintenance of bear-

ings usually involves periodicregreasing (followed by a heat run)and monitoring bearing vibrations,which will gradually increase overlong periods of time .

To maintain pumps and dri-vers that are secured for longperiods of time, simply turn therotor 10–15 revolutions everythree months by hand. This willensure than an adequate greasefilm exists to prevent corrosion ofthe bearing. If this action is nottaken, the bearings may begin tocorrode due to a breakdown inthe grease film. ■

Ray W. Rhoe, PE, has a BSCEfrom The Citadel and 15 years’ expe-rience with pumps, testing, andhydraulic design.

Zero-leakage magnetic liquid seal developed to retrofit process pumps

Page 30: Centrifugal Pumps Handbook

30 The Pump Handbook Series

BY DAVID CUMMINGS

he efficiency of centrifugalpumps of all sizes is becomingmore important as the costand demand for electricity

increases. Many utilities are empha-sizing conservation to reduce thenumber of new generating facilitiesthat need to be built. Utilities haveincreased incentives to conservepower with programs that emphasizedemand side conservation. These pro-grams often help fund capital equip-ment replacements that reduceelectrical consumption. Demand sidemanagement programs make replac-ing old pumping equipment morefeasible than ever before.

DETERMINING PUMP EFFICIENCYThe efficiency of a pump ηp is

ratio of water horsepower (whp) tobrake horsepower (bhp). The highestefficiency of a pump occurs at theflow where the incidence angle of thefluid entering the hydraulic passagesbest matches with the vane angle.The operating condition where apump design has its highest efficien-cy is referred to as the best efficiencypoint (BEP).ηp = whp/bhp

The water horsepower (whp)may be determined from the equa-tion:whp = QHs/3,960where Q = capacity in gallons perminute, H = developed head in feet,and s = specific gravity of pumpedfluid.

Preferably, the brake horsepow-er supplied by a driver can be deter-mined using a transmissiondynameter or with a specially cali-brated motor. Brake horsepowerdetermined in the field by measuringkilowatt input and multiplying by themotor catalogue efficiency can beinaccurate. If motor power is deter-mined in the field, data should betaken at the motor junction box, notat the motor control center.

Overall pump motor efficiencyηo may be determined from the equa-tion:ηo = whp/ehp

T where ehp = electrical power inhorsepower.

EFFICIENCY LOSSESPump efficiency is influenced by

hydraulic effects, mechanical losses,and internal leakage. Each of thesefactors can be controlled to improvepump efficiency. Any given designarrangement balances the cost ofmanufacturing, reliability, and powerconsumption to meet users needs.

Hydraulic losses may be causedby boundary layer effects, disruptionsof the velocity profile, and flow sepa-ration. Boundary layer losses can beminimized by making pumps withclean, smooth, and uniform hydraulicpassages. Mechanical grinding andpolishing of hydraulic surfaces, ormodern casting techniques, can beused to improve the surface finish,decrease vane thickness, andimprove efficiency. Shell molds,ceramic cores, and special sands pro-duce castings with smoother andmore uniform hydraulic passages.

Separation of flow occurs when apump is operated well away from thebest efficiency point (BEP). The flowseparation occurs because the inci-dence angle of the fluid entering thehydraulic passage is significally dif-ferent from the angle of the blade.Voided areas increase the amount ofenergy required to force the fluidthrough the passage.

Mechanical losses in a pump arecaused by viscous disc friction, bear-ing losses, seal or packing losses, andrecirculation devices. If the clearancebetween the impeller and casing side-wall is too large, disc friction canincrease, reducing efficiency.Bearings, thrust balancing devices,seals, and packing all contribute tofrictional losses. Most modern bear-ing and seal designs generate fullfluid film lubrication to minimize fric-tional losses and wear. Frequently,recirculation devices such as auxiliaryimpellers or pumping rings are usedto provide cooling and lubrication tobearings and seals. Like the mainimpeller, these devices pump fluid

and can have significant powerrequirements.

Internal leakage occurs as theresult of flow between the rotatingand stationary parts of the pump,from the discharge of the impellerback to the suction. The rate of leak-age is a function of the clearances inthe pump. Reducing the clearanceswill decrease the leakage but canresult in reliability problems if matingmaterials are not properly selected.Some designs bleed off flows fromthe discharge to balance thrust, pro-vide bearing lubrication, or to coolthe seal.

EXPECTED EFFICIENCIESThe expected hydraulic efficien-

cy of a pump design is a function ofthe pump size and type. Generally,the larger the pump, the higher theefficiency. Pumps that are geometri-cally similar should have similar effi-ciencies. Expected BEPs have beenplotted as a function of specific speedand pump size. A set of curves thatmay be used to estimate efficiency isprovided in Figure 1. The specificspeed (Ns) of a pump may be deter-mined from the equation:Ns = NQ0.5/H0.75

where N = speed in rpm, Q = capac-ity in gpm, and H = developed headin feet.

Using a pump performancecurve, the highest efficiency can bedetermined and the specific speedcalculated using the head and capaci-ty at that point. Using the specificspeed and the pump capacity, theexpected efficiency can be estimated.If the pump has bearings or seals thatrequire more power, such as tiltingpad thrust bearings or multiple seals,this should be considered when cal-culating the expected efficiency.

MOTOR EFFICIENCIESEfficiencies for new “premium

efficiency motors” are provided inFigure 2. Using these values, withanticipated pump efficiencies inFigure 1, the expected power con-sumption for a well designed pump

Centrifugal Pump Efficiency

CENTRIFUGAL PUMPS

HANDBOOK

Page 31: Centrifugal Pumps Handbook

The Pump Handbook Series 31

and motor can be determined. Thecalculated power consumption can becompared with an existing installa-tion to determine the value ofimproving pump performance orreplacing the unit.

EXAMPLE CALCULATION OFPUMP EFFICIENCY

A single-stage end-suctionprocess pump will be used as anexample for an efficiency calculation.The pump uses a mechanical seal andan angular contact ball bearing pairfor thrust. The pumped fluid is waterwith specific gravity of 1.0. Thepump operates at its BEP of 2,250gpm, developing 135 feet of head.The pump speed is 1,750 rpm (note:with the new motor the speed maychange, but to simplify the example itwill be assumed the new and oldmotor both operate at 1,750 rpm).The expected power consumption fora new unit can be calculated.

First the pump specific speedwill be calculated:

[1,750 rpm x (2,250 gpm)0.5]Ns = (135 ft)0.75

Ns = 2,096Figure 1 can be used at Ns = 2,100and interpolated for 2,250 gpm.The expected efficiency is 86%.The water horsepower is:

(2,250 gpm x 135 ft x 1.0)whp = 3,960

whp = 76.7 HpThe expected brake horsepower is:bhp = 76.7 Hp/0.86bhp = 89.2 Hp

The antifriction bearings and typ-ical mechanical seal do not require apower adjustment. However, if a tilt-ing pad thrust bearing or otherdevice, such as a special seal, thatused more power was part of thedesign, the correction would be madehere by adding the additional horse-power to the calculated value.

Using Figure 2, the minimumefficiency for a 100 Hp motor is 95%.The efficiency value may changeslightly for the operating conditionand should be verified with a motormanufacturer. The 95% efficiencywill be used in this example. Theexpected electrical horsepower is:ehp = 89.2 Hp/0.95ehp = 93.9 Hp

orehp = 93.9 Hp x.7457 kw/Hpehp = 70.0 kwThe last time thispump was rebuiltand put in ser-vice, power wasmeasured at 79.5kw. The differ-ence in powerc o n s u m p t i o nbetween the exist-ing unit and anew unit can becalculated:ehp = 79.5 kw -70.0 kwehp = 9.5 k.

If power costs 8 cents a kilowatt hourand the pump operates continuously,the savings of replacing this unit onan annual basis can be calculated:cost = 9.5 kw x $0.08 per kw hr x8,760 hr/yrcost = $6,658 year

This figure can be used to deter-mine if the additional power con-sumption justifies replacing thepump. If a replacement pump andmotor of this size can be purchasedand installed for $40,000, and theelectric utility offers a 50% rebateprogram, the net cost of $20,000 forthe user is certainly worth consider-ing.

SUMMARYRemember, for any centrifugal

pump to operate efficiently it needsto be properly applied. Whenprocesses require changing flow ratesfrequently, variable speed drives canbe a solution. A pump operating farfrom its BEP will be neither efficientnor reliable. Many times changingthe pump size to better match thesystem will reduce power costs dra-matically. ■

David L. Cummings is an indepen-dent consultant who provides engineer-ing services and equipment forspecialized applications.

Motor Minimum Acceptable EfficiencyHorsepower 1200 rpm 1800 rpm 3600 rpm5 88.0 88.0 87.010 90.2 91.0 90.215 91.0 92.0 91.020 91.7 93.0 91.725 92.4 93.5 92.030 93.0 93.6 92.440 93.6 94.1 93.050 93.6 94.1 93.075 94.5 95.0 94.1100 94.5 95.0 94.5125 94.5 95.4 94.5150 95.0 95.4 94.5200 95.0 95.0 95.0Over 200 95.0 95.4 95.0

TABLE 1. EXPECTED EFFICIENCY FOR “PREMIUM EFFICIENCY MOTORS”

FIGURE 1

Efficiency of various pumps sizes and specific speeds

Page 32: Centrifugal Pumps Handbook

32 The Pump Handbook Series

How do I select theproper motor size formy centrifugal pumpapplications? In some

applications we have experi-enced driver overload, whileother applications appear satis-factory using the same selec-tion method.

What seems to be astraightforward re-quirement of selectinga pump motor too fre-quently results in a

major problem when commission-ing the pumping system and theinstalled motor overloads and istripped off the line. Correcting theproblem can be as simple asadjusting a manual valve. And itcan be as complex and time con-suming as replacing the motor,motor starters, and service wirewith a larger size, and altering theflow control system.

A simple rule of thumb ofsupplying a motor size thatexceeds the pump manufacturersrated point brake horsepower bysome fixed margin, or of supply-ing a motor size equal to the brakehorsepower at the end of theselected impeller diameter curve,may not work in all cases. Theperson selecting the motor musthave a thorough knowledge of thepumping system and its character-istics, pump industry practices,and limitations of the generalizeddata provided by the pump manu-facturer with the quotation.

GENERALIZED PERFORMANCECURVE LIMITATIONS

The brake horsepower valuespublished by the pump manufac-turer on the generalized hydraulicperformance curves (TDH, effi-ciency, BHP vs. capacity) are thebasis for the rated point BHPreturned to a potential customerwhen responding to a request forquotation for a specific applica-tion. The data is as accurate as

practical for the designated equip-ment design, but does have a toler-ance range which may be ex-perienced for any specific pump forsuch characteristics as the TotalDynamic Head (TDH) at a specificflow rate.

Brake horsepower is related tothe TDH by the following formula:

BHP =(TDH) x Flow x Specific Gravity

3960 x Efficiency x Viscosity Correction Factors*

*See Hydraulic Institute Standards for Values

Most pump users will not accepta lower than specified rating pointTDH, and the manufacturer is fre-quently required to test the equip-ment prior to shipment to confirmthat the pump meets the specifiedrequirements. If the installedimpeller should produce less TDHthan specified, the manufacturermust replace the impeller with alarger diameter. If high alloy materi-als are involved, there may be con-siderable expense and delayinvolved. Hence the practice in thepump industry is to publish a perfor-mance curve (TDH vs. Capacity) fora given impeller diameter that issomewhat less than can actually beachieved by the specified diameters.Should a test then be specified andthe impeller TDH test higher thanallowed tolerances at a given flowrate, the impeller can be reduced to asmaller diameter to provide therequired values without replacing it.

One of the current pump indus-try acceptance test criteria, theHydraulic Institute Standards, per-mits the TDH to exceed the designpoint requirements by as much as8%. Sometimes pump impellers willexceed the published data by asmuch as 20% when first tested. Thisis not usually the case, but there is arange of performance, especially on

relatively new or infrequently test-ed pump designs.

The user should be informed of this potential variation if theimpeller requires replacementdue to normal operating wear ofthe pump, especially if it is to bepurchased from a source otherthan the original equipment man-ufacturer, which may not havehistorical test records of the origi-nal hydraulic design.

Industry practice is to guaran-tee only the TDH (with a tolerancerange) at a specified flow rate andthe pump efficiency. The resultingbrake horsepower is guaranteedonly by the same tolerance, andthen only if the pump is tested.

MECHANICAL SEAL HORSEPOWER LOSSES

The performance curve pub-lished by the pump manufacturerand described above does not pro-vide allowances for the powerrequired to turn a mechanical sealthat is loaded to typical processconditions. For a high suctionpressure, double mechanical shaftseal pump installation, this can bea measurable amount and must beadded to the horsepower requiredto move the liquid. An allowanceof one to two horsepower, forexample, may be required forsome ANSI style pump designs tocompensate for seal losses. Hence,if the generalized performancecurve rating point results in a BHPof 7.5 Hp, a motor of 10 Hp maybe considered for the application ifno allowance is given for factorslike seal drag.

As part of the equipment quo-tation, an estimate for the sealhorsepower drag should berequested for all pumps requiringmechanical seals. If a doublemechanical seal has been specifiedwith a buffer fluid pressurizedflush, the buffer fluid pressuremust be specified by the seal andpump manufacturer and observed

Motor Size Selection for Centrifugal PumpsBY ROBERT J. HART

Q:

A:

CENTRIFUGAL PUMPS

HANDBOOK

Page 33: Centrifugal Pumps Handbook

The Pump Handbook Series 33

by the user to assure the estimatedseal drag horsepower is notexceeded. Over pressurizing thedouble mechanical seal buffer sys-tem at the site can result in amotor overload condition notanticipated during the motor selec-tion phase.

Seal horsepower losses typi-cally have a greater impact on theinstallations at or below 25 Hp,but they should be considered forall installations.

FLOW CONTROLSince the BHP of most pump

designs increases with increasingflow through the pump, it is theuser’s responsibility to assure thatthe actual system flow does notexceed the rated flow originallyspecified when the pump was pur-chased. Pumping systems thatlimit flow only by the resistance ofinstalled piping have a tendency tobe sized with safety factors to“assure” the pump selected willprovide adequate flow (see theabove comments regarding gener-alized performance curves). Amotor may overload when thepump operates at a higher flowrate than anticipated and requiresa greater horsepower. Should theactual system curve extendbeyond the end of the publishedpump curve and not intersect thepump curve, the actual horsepowerwill be greater than the “end ofcurve” horsepower frequentlyused as basis for motor selection.With adequate NPSH available tothe pump, the performance curveand corresponding horsepowermay extend to greater than pub-lished values (Figure 1).

FLUID CHARACTERISTICSBoth specific gravity and vis-

cosity can affect the requiredpump brake horsepower (seeequation above). Motors are nor-mally selected on the basis ofrated conditions of head, flow,specific gravity, temperature, andviscosity.

The off-design conditions ofthese characteristics should be

examined and thosefluid characteristicswhich affect brakehorsepower evaluatedbefore selecting a dri-ver.As examples:

Is there an alter-nate start-up or shut-down flush liquidrequired which has ahigher specific gravityliquid than the ratedflow material?

What is the actu-al liquid viscosity at alower temperaturethan rated conditions,and will it increasethe BHP of thepump? Even thoughthe pump and pipingis well insulated,without heat tracingthe system will be atambient temperatureduring a start-up. Thiswill cool the incom-ing liquid below thecontinuous on-line conditions thatwould exist once the piping systemis in operation and at equilibrium.

PUMP WEARA certain amount of internal

recirculation takes place inside acentrifugal pump casing at all times.As internal clearances change due towear, the rate of this circulationincreases. If the system demandsdown stream of the pump remainconstant and the system is designedto maintain process flow, the pumpmust flow at a higher rate to com-pensate for this recirculation.Because of this, it may require a cor-responding higher horsepower. SeeI.J. Karassik’s recent article “Whento Maintain Centrifugal Pumps”(Hydrocarbon Processing, April1993) for additional information onthis topic.

SUMMARYMotor selection for centrifugal

pumps involves many considera-tions, some of which are beyond the

control of the pump manufacturer.There is no simple rule of thumb.

Oversizing motors to compen-sate for all of the conditions thatmay or may not exist on everyinstallation can be a major addi-tional expense when consideringthe total electrical system. Thisarticle has illustrated the variablesthat must be taken into account. ■

Robert J. Hart is a SeniorConsultant at the DuPont Company.He also serves on the Pumps andSystems Editorial Advisory Board.

A

B

TDH

BHP

Flow

FIGURE 1

Pump performance curve. A=calculated systemcurve with safety factors, B=actual system curve.

Page 34: Centrifugal Pumps Handbook

34 The Pump Handbook Series

strong controversy. Accept my com-ments as a personal opinion.

One theoretical method exists topredict the onset of recirculation (Ref.1 and 2). The results of this methodhave been verified by many tests,with actual pumps and plastic trans-parent models where the onset couldbe observed with a strobe light. Theresults corresponded within no morethan 5% deviation from the predic-tions.

Assuming that the pump is prop-erly furnished with the necessaryinstrumentation, such as flowmeter,pressure gauges with sufficient sensi-tivity to show pulsations, and vibra-tion and noise monitoring devices, anexperienced test engineer should beable to pinpoint the onset of internalrecirculation.

But it is the setting of the mini-mum flow which—for the timebeing—remains controversial.Obviously, any material damage can-not serve as a standard, because bythe time the correctness of the deci-

I am presently in-volved in replacing anewly purchased pump.It was accepted by thepurchaser, but the

shop test was noisy. The manufac-turer said this was due to the poorsuction piping. The field test wasunacceptable and noisy, and therewas disagreement about whetherthe noise was due to improperlyplaced elbows in the suction pipingor if the pump was inappropriatelyselected. The pump, probablydesigned for flows much greaterthan system requirements, wasrecirculating. The noise was a verylow frequency, random banging.The single-stage, double-suction,twin-volute design had four timesmore NPSH than required. Howwould your experts have diagnosedthis costly problem?

What witness shop test shouldbe conducted so that the pumppurchaser can be assured of a safecontinuous flow as quoted in theproposal? What measurements,observations (both audible andvisual), and instrumentationshould be used to detect the onsetof suction and/or discharge recircu-lation? If the pump does not per-form as quoted, are minor shopalterations conceivable?

I would like to suggest to theHydraulic Institute that a minimumnonrecirculating flow test be addedto the standards. Your thoughts?

J. P. Messina, ProfessionalEngineer, Pump and HydraulicsConsultant, Springfield, NJ.

Until about 25 years ago,there were only four fac-tors to consider when set-ting an acceptableminimum flow for cen-

trifugal pumps:

• higher radial thrust developedby single volute pumps atreduced flows

• temperature rise inthe liquid pumped

• desire to avoidoverload of driversof high specific-speed axial-flowpumps

• for pumps hand-ling liquids withsignificant amountsof dissolved orentrain-ed air orgas, the need tomaintain sufficient-ly high fluid veloci-ties to wash out thisair or gas alongwith the liquid

Since then, a newphenomenon has beendiscovered that affectsthe setting of minimumflows. At certainreduced flows, all cen-trifugal pumps are sub-ject to internalrecirculation, both in the suction anddischarge areas of the impeller. Thisproduces pulsations at both the suc-tion and discharge, and the vibrationcan damage impeller material in away similar to classic cavitation,although taking place in a differentarea of the impeller.

Each of these effects may dictatea different minimum operatingcapacity. Clearly, the final decisionmust be based on the greatest of theindividual minimums. The internalrecirculation usually sets the recom-mended minimum, which appears tobe what happened in your case.

You’ve actually raised twoquestions:1. How can one determine the

onset of recirculation?2. After determining the onset,

what should be the recommend-ed minimum flow in relation tothe recirculation flow?

Unfortunately, the answers toboth questions remain in the realm of

Setting the Minimum Flows for Centrifugal PumpsBY: IGOR J. KARASSIK

Q:

A:

Unstableregionpump B

HRSA

QRSA

HRSB

QSRB

25%

Safe Zonesof Operation

Capacity of Q in%

Head

, H

Min

. Flo

w

Min

. Flo

w

Pum

p A

Pum

p B

100%

FIGURE 1

Comparison of safe zones of operation fornormal and for high S value impellers

CENTRIFUGAL PUMPS

HANDBOOK

Page 35: Centrifugal Pumps Handbook

The Pump Handbook Series 35

sion has been verified, it is too late.Therefore, the magnitudes of pres-sure pulsation, noise and vibrationare the only criteria for establishingthe minimum flow.

Regarding vibration, theHydraulic Institute Standardsincludes a chart, plotting maximumpermissible peak-to-peak amplitudesagainst frequency, and it is applica-ble “when the pump is operating atrated speed within plus or minus10% of rated capacity.” This couldcreate a serious problem whenever apump meets these limits but is sub-ject to considerably higher vibrationswhen operated below the recircula-tion flows. The API-610 Standard ismore specific, defining the minimumcontinuous stable flow at which thepump can operate without exceedingthe noise and vibration limitsimposed by the Standard. These lim-its are expressed in inches per sec-ond rather than mils ofdisplacement.

The Hydraulic In-stitute shouldprobably set up rules for establishinga minimum flow test. But obtaining

a consensus about theacceptable limits ofvibration and noise willbe difficult. The choiceof a minimum flow ismuch more subjective ifit is based on problemsarising from internalrecirculation than whenthe temperature rises,and radial thrust andoverload of drivers of high specific-speedpumps are concerned.In these situations, theeffect of operating atany given low capacitycan be quantified. Eventhe effect of handlingliquids laden with air orgas is easy to determinesince at some givenflow the noncondensi-ble content of the liquidwill not be washedaway, and will accumu-late within the pump,which will stop pump-ing.

Another majorobstacle to overcome in

achieving a consensus is to definewhat is continuous service and whatis intermittent. When Warren Fraser(who did all the seminal work oninternal recirculation) and I tried toproduce a quantitative value thatwould distinguish between these two,we first tried to define 25% of thetime as the breaking point betweenthem. At first this seemed reasonable,but we soon realized that we hadanother problem to face. There was adifference between running a pumpfor six hours per day at or below anarbitrary flow and running it for threemonths out of a year for some strictlyclimatic conditions. So, we decided toavoid making any formal distinctionbetween continuous and intermittent.

I admit that I do not have a defi-nite and final answer to offer on thesubject of selecting a minimum flowstandard. I continue to use a guide-line that I established some years ago(Ref. 3). Because the choice ofrequired NPSH affects the onset ofinternal recirculation, for high suctionspecific speeds the minimum flowshould correspond to the onset of the

recirculation. For cold water, thisrefers to values over 10,500. And formore conservative S values, such as8500 to 9500, set the minimum flowat 25% of the best efficiency capacity(Figure 1).

These comments represent mypersonal opinion. I am aware thatsome users may be more conservativeand insist that the minimum flowshould never be less than the recircu-lation capacity. In that case, usersshould specify this restriction in their

FIGURE 2

“Bulk-head ring” construction used to elimi-nate unfavorable effects of excessively largeimpeller eye diameter

FIGURE 3

Projections from casing wallprovided to reduce problemscreated by discharge siderecirculation

FIGURE 4

Addition of two annular ringsto impeller shrouds to reduceaxial movement of rotorcaused by internal recircula-tion at discharge

Projectionsfrom casingwall to reduceaxialunbalance

Annular ringAnnular ring

Page 36: Centrifugal Pumps Handbook

the ASME Winter AnnualMeeting (1981).

3. I.J. Karassik. Centrifugalpump operation at off-designconditions. Chemical Processing (1987).

Igor J. Karassik is SeniorConsulting Engineer for Ingersoll-Dresser Pump Company. He hasbeen involved with the pump industryfor over 60 years. Mr. Karassik is amember of the Pumps and SystemsEditorial Advisory Board.

you can achieve some relief byproviding projections from thecasing wall (Figure 3).Alternately, annular rings can befitted to the outer shrouds of theimpeller (Figure 4).

I hope these comments will serveto open a dialogue between pumpusers and manufacturers. Such a dis-cussion should lead to the undertakingof a series of tests that will shed addi-tional light on the problem of accept-able minimum flows. These tests, inturn, could permit the HydraulicInstitute to include guidelines in itsstandards. ■

REFERENCES

1. W.H. Fraser. Flow recirculationin centrifugal pumps. Proce-edings of the Texas A&M TenthTurbomachinery Symposium(1981).

2. W.H. Fraser. Recirculation incentrifugal pumps. Presented at

requests for bids. But this couldonly be acceptable if guidelinesare published on how to conduct atest for recirculation or a formulabecomes widely accepted on howto calculate the onset of internalrecirculation.

Regarding your questionabout what minor alterations maybe made if the pump does not per-form satisfactorily in this connec-tion, there are two possiblesolutions:1. For suction recirculation, you

can reduce the minimumacceptable flow by incorporat-ing a “bulk-head ring” with anapron overhanging the eye ofthe impeller (Figure 2). Ofcourse, this does increase therequired NPSH and can onlybe done if there is the neces-sary margin between avail-able and required NPSH.

2. If the problem is caused bydischarge side recirculation,

36 The Pump Handbook Series

Page 37: Centrifugal Pumps Handbook

The Pump Handbook Series 37

BY JAMES NETZEL

How can you estimatethe maximum (shutoff)head that a centrifugalpump can deliver?

The maximum pressure acentrifugal pump deliversshould be known in orderto ensure that a pipingsystem is adequately

designed. Any pump that operatesat a high flow rate could deliversignificantly more pressure at zero(0) gpm flow, such as when the dis-charge valve is closed, than itdelivers at operating flow.

The maximum head or dischargepressure of a centrifugal pump,which usually occurs at shutoff con-

ditions (0 gpm), can be easily esti-mated if the impeller diameter,number of impellers used, andrpm of the driver (electric motor,gas engine, turbine, etc.) areknown.

Let’s say we have a single-stage pump with a 10-in. diameterimpeller and an 1,800 rpm driver.To determine the head in feet,simply take the impeller diameterin inches and square it. Our 10-in.impeller at 1,800 rpm would yield102, or 100 ft of head. An 8-in.impeller would yield 82, or 64 ft ofhead, while a 12-in. impellerwould yield 122, or 144 ft of head.

Now let’s assume that our10-in. diameter impeller is drivenby a 3,600 rpm motor. We firstdetermine the head at 1,800 rpm,but then multiply this value by afactor of four. The basic rule isthat every time the rpm changesby a factor of two, the headchanges by a factor of four. Thehead at 3,600 rpm for our 10-in.impeller is therefore 102 x 4, or400 ft of head. Our 8-in. impellerat 3600 rpm would give us 82 x 4,or 256 ft of head, and our 12-in.impeller would give us 122 x 4, or576 ft of head.

For multiple stages (morethan one impeller), simply multi-ply the final head for one impellerby the total number of impellersin the pump. For a pump withthree 10-in. impellers and a speedof 3,600 rpm, we get (102 x 4) x3 = 400 x 3 = 1,200 ft. of head.

Now what happens if wereduce the speed below 1,800 rpm?The same rule still applies: achange in speed by a factor of twochanges the head by a factor offour. Therefore, a 10-in. diameterimpeller spinning at 900 rpm deliv-ers only one fourth the head itwould at 1,800 rpm: 102/4 = 25 ft.

Plotting several head-versus-rpm points on a curve will allowthe user to estimate the maximum

The maximum head or discharge

pressure of a centrifugal pump can be

easily estimated if the impeller

diameter, number of

impellers used, and rpm

of the driver are known.

Estimating Maximum Head in Single – and Multi-Stage Pump Systems

Q:A:

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15

1716

15

14

13

121110

9

8

7

65432

1

Hea

d in

Fee

t x 1

000

RPM x 1000

FIGURE 1

Rotations per minute (rpm) vs. head in feet to estimate maximum head

CENTRIFUGAL PUMPS

HANDBOOK

Page 38: Centrifugal Pumps Handbook

38 The Pump Handbook Series

What different types ofseal lubrication exist?

A mechanical seal isdesigned to operate inmany types of fluids. Theproduct sealed becomesthe lubricant for the seal

faces. Many times the fluid beingsealed is a poor lubricant or containsabrasives that must be taken intoaccount in the seal design. Thedesign of the seal faces, materials ofconstruction, and seal lubricationplay an important role in successfuloperation. Achieving a high level ofreliability and service life is a clas-sic problem in the field of tribolo-gy, the study of friction, wear, andlubrication.

The lubrication system for twosliding seal faces can be classified asfollows: 1) hydrodynamic, 2) elasto-hydrodynamic, 3) boundary, and 4)mixed film.

Hydrodynamic conditions existwhen the fluid film completely sepa-rates the seal faces. Direct surfacecontact between seal faces does nottake place, so there is no wear, andheat generation from friction is zero.The only heat generation occursfrom shearing of the fluid film,which is extremely small. A hydro-dynamic seal may rely on design fea-tures such as balance factors, surfacewaviness, or spiral grooves to sepa-rate the seal faces. The Society ofTribologists and LubricationEngineers (STLE) guideline in“Meeting Emissions Regulationswith Mechanical Seals” lists hydro-dynamic seals as a technology tocontrol emissions.

Elastohydrodynamic lubrication(EHD) is also found in sliding sur-faces, but more often this involvesrolling surfaces separated by an oilfilm. Here the moving surfaces form

head at any given speed. Let’s saywe have a turbine-driven pumpthat injects water into the groundto raise the subterranean oilreserves to the surface for process-ing. The vendor tells you that themaximum head is classified, butyou have been requested toresolve system problems that youbelieve are pressure related. Thevendor tells you that the pumphas four 8-in. diameter impellersand is driven by the turbine at13,000 rpm. You would estimatethe maximum head as follows:

Step 1 Determine the head at1,800 rpm:82 x 4 stages = 256 ft

Step 2 Multiply the head at 1,800 rpm by four to getthe head at 3,600 rpm:256 x 4 = 1,024 ft

Step 3 Multiply the head at 3,600 rpm by 4 to get thehead at 7,200 rpm:1,024 x 4 = 4,096 ft

Step 4 Multiply the head at 7,200 rpm by 4 to get thehead at 14,400 rpm:4,096 x 4 = 16,384 ft

Step 5 Plot the rpm-versus-headpoints to obtain the curveshown in Figure 1.

As you can see, the estimatedhead at 13,000 rpm is 12,500 ft. Toconvert head in feet to psi, simplydivide the head by 2.31 to get5,411 psi.

Ray W. Rhoe, PE, has a BSCEfrom The Citadel and 15 years’ expe-rience with pumps, testing, andhydraulic design.

Q:an interface region that deformselastically under contact pressure.This deformation creates largerfilm areas and very thin films.Such lubrication systems are nor-mally used to control wear inrolling element bearings. In sealswhere the viscosity of the fluidsealed increases with increasingpressure,elastohydrodynamiclubrication occurs.

Boundary lubrication isimportant for seal faces that aremoving very slowly under heavyload. Here, hydrodynamic andelastohydrodynamic lubricantpressures are insufficient to sepa-rate the seal faces. The slidingsurfaces are protected by the tri-bological properties of the materi-als of construction. An exampleof a seal operating within thislubrication system is a dry-runningagitator seal.

Mixed-film lubrication, a com-bination of all the previous sys-tems discussed, occurs in allcontact seals. Here the fluid filmbecomes very thin and is a combi-nation of both the liquid and thegas phases of the fluid sealed.Asperities from one surface maypenetrate the lubricating film andcontact the opposite surface. Theseal face load is then supportedpartially by the fluid film and par-tially by solid contact. If the gener-ated head at the seal faces is notremoved, surface wear and dam-age can occur. For applicationswhere the seal face load is toohigh or the fluid viscosity is toolow, designs of seal faces can bechanged through balance and facegeometry to improve seal perfor-mance. ■

James Netzel is Chief Engineerat John Crane Inc. He serves on theEditorial Advisory Board for Pumpsand Systems.

A:

Page 39: Centrifugal Pumps Handbook

The Pump Handbook Series 39

BY WILLIAM E. (ED) NELSON

When trimming a pumpimpeller to change the flowand head, I sometimes gettoo much of a reduction.What is the problem?

With a constant rotationalspeed, as is the case withmost pumps, the “AffinityLaws” commonly used forcalculating the trim do not

accurately reflect the relationshipbetween the change in impeller diam-eter and the hydraulic performanceachieved by the pump. The calcula-tions generally dictate more of a cutthan required to affect the desiredhead and flow reduction. The“Affinity Law” errors can be on theorder of 20 percent of the calculatedreduction. If the calculated reductiontrimming calls for a 10 percent reduc-tion in diameter, only seven or eightpercent reduction should be made.The lower the specific speed of theimpeller cut, the larger the discrepan-cy. This subject is covered in only afew pump handbooks. The subject iswell covered on pages 18 and 19 ofCentrifugal Pumps - Design andApplication, First Edition, by ValLobanoff and Robert R. Ross. Thereare several reasons for the actualhead and flow being lower than thatcalculated:

1.The “Affinity Laws” assumethat the impeller shroudsare parallel. In actuality, theshrouds are parallel only inlower specific speed pumps.

2.The liquid exit angle is alteredas the impeller is trimmed, sothe head curve steepens slightly.

3.There is increased turbulentflow at the vane-tips as theimpeller is trimmed, if theshroud-to-casing clearance(Gap “A”) is not maintained.

All of these effects contribute toa reduced head development andflow. Pumps of mixed flow designare more affected than the true radialflow impellers found in higher headpumps. More caution has to be exer-

Q:A:

cised in altering the diameter of amixed flow impeller.

What is the effect oftrimming an impeller onpump efficiency?

It depends on the specificspeed of the impeller. Thespecific speed index classi-fies the hydraulic featuresof pump impellers accord-

ing to their type and proportions.Most refinery pumps fall betweenabout 900 and 2,500 on this index.Some vertical multistage pumps arein the 4,000 to 6,000 range.

For radial designs, impellerdiameter should not be reduced morethan 70 percent of the maximumdiameter design. Reductions in pumpimpeller diameters also alter outletchannel width, blade exit angle andblade length and may significantlyreduce the efficiency. The greater theimpeller diameter reduction frommaximum diameter and the higherthe specific speed (not suction specificspeed), the more the pump will

decrease with the trimming of theimpeller.

What is the effect ofimpeller trimming onNPSH required?

Small reductions inimpeller diameter willincrease the requiredNPSH only slightly.Diameter reductions great-

er than about five to 10 percent willincrease NPSH required, whichoccurs because specific vane loadingis raised by the reduced vane length,affecting velocity distribution at theimpeller inlet. Not all pump compa-nies consistently show their pumpcurves the increase NPSHR withreduced impeller diameters. Atten-tion must be paid to this factor whenthe margin between NPSHR andNPSHA is very narrow or the NPSHRfor a pump is extremely low.

What effect does trim-ming an impeller have onaxial vibration?

Excessive impeller shroud-to-casing clearances (Gap“A”) and suction recircula-tion cause eddy flowsaround the impeller, which

in turn cause low frequency axialvibrations. Flow disturbances relatedto suction recirculation and cavitationare always present in both diffuserand volute type pumps. As the

Tips on Pump Efficiency

FIGURE 1. SHROUD & VANE REDUCTION

FIGURE 2. OBLIQUE CUTS OF VANE

CENTRIFUGAL PUMPS

HANDBOOK

Q:A:

Q:A:

Q:A:

GAP “A”

GAP “B”

(a)

D’ D

Gap “A”

Gap “B”

Gap “A”

Gap “B”

D’ D

D’ D

Page 40: Centrifugal Pumps Handbook

40 The Pump Handbook Series

impeller diameters are reduced, theflow distribution pattern across theexit width of the impeller becomesmore unstable. The tendency for thehigh-pressure liquid to return to thelow pressure side and create tip recir-culation is greatly increased. Again,the higher energy level pumps are ofmajor concern (above 200 HP and650 feet of head per stage).

What are the effects oftrimming an impeller onradial vibration?

Careful machining of thevolute or diffuser tips toincrease Gap “B” whilemaintaining Gap “A” hasben used for a number of

years to greatly reduce the vane-pass-ing frequency vibration. The pulsat-ing hydraulic forces acting on theimpeller can be reduced by 80 to 85percent by increasing the radial Gap“B” from 1 percent to 6 percent.There is no loss of overall pump effi-ciency when the diffuser or voluteinlet tips are recessed, contrary to theexpectations of many pump design-ers. The slight efficiency improve-ment results from the reduction ofvarious energy-consuming phenome-na: the high noise level, shock, andvibration caused by vane-passing fre-quency, and the stall generated at thediffuser inlet.

Table 1 gives recommendeddimensions from Dr. Elemer Makayfor radial gaps of the pump impellerto casing. Note that if the number ofimpeller vanes and the number of dif-fuser/volute vanes are both even, theradial gap must be larger by about 4percent.

When trim-ming an im-peller fromits maximumdiameter to

adjust the head andflow developed by acentrifugal pump,what is the best wayto cut the impeller? Isit best to trim theimpeller vanes andthe shrouds or just thevanes?

No hard andfast guidelinesfor the mechan-ical aspects ofimpeller trim-

ming exist, but there are severalpump construction and hydraulicdesign factors to consider while mak-ing the decision of what to trim.

How the impeller is trimmedwill greatly influence the hydraulicperformance of the pump as well asthe vibration levels experienced. Youmust evaluate the hydraulic charac-teristics before you decide how totrim the impeller.

For volute type pumps, theentire impeller, vanes and shroudsmay be cut as shown in Figure 1.However, in some pumps, thismethod will alter Gap “A” (shroud-to-case clearance), leading to unevenflow distribution at the impeller exitarea, which can cause axial vibrationand other problems. The double suc-tion impeller type pump is especiallysensitive to problems caused byincreased Gap “A”, so trimming theentire impeller is not a good choice. It

is best to cut the impeller vanesobliquely (Figure 2), which leaves theshrouds unchanged or to cut thevanes only (Figure 3). Trimming thevanes only tends to even out the exitflow pattern and reduce recirculationtendencies at the exit area. Gap “A”should be about 0.050 inch (radial)for minimum vibration due to vane-passing frequency.

In most diffuser type pumps, it isbest to trim only the vanes (Figure 3)to control tip recirculation and the illeffects of an increased Gap “A”. Thiscut yields a more stable head curve.The uniform flow reduces the ten-dency for tip recirculation, and thepossibility of suction recirculation isgreatly reduced at the exit area.

Structural strength of the shroudsis a factor in the decision in how totrim the impeller. There may be toomuch unsupported shroud left after amajor reduction in diameter. The

FIGURE 3. TERMINATING VANES ONLY

Type of Gap “A” Gap “B” +/- percentagePump Design of impeller radius

Minimum Preferred MaximumDiffusers 50 mils 4% 6% 12%Volute 50 mils 6% 10% 12%

*B = 100 (R3-R2)R2

where R3 = Radius of diffuser of volute inletand R2 = Radius of impeller

TABLE 1. RECOMMENDED RADIAL GAPS FOR PUMPS

FIGURE 4. IMPELLER VANE OVERFILING

Length of blend for over filingImpeller diameter, in “A” distance of blend, in10 & below 1 1/2

10 1/16 through 15 2 1/2

15 1/16 through 20 3 1/2

20 1/16 through 30 530 & larger 6

Q:A:

Q:

A:

Page 41: Centrifugal Pumps Handbook

The Pump Handbook Series 41

I frequently encounter“vane-passing” frequen-cies during vibrationanalysis of a pump.What are some of the

methods that can be used toreduce this problem?

The most effective methodof reducing vane-passingfrequencies is to carefullymaintain proper Gap “A”and Gap “B” clearances to

reduce impeller-casing interaction.Sometimes, impellers manufacturedwith blunt vane tips cause distur-bances in the impeller exit area andin the volute area by generatinghydraulic “hammer” even when theimpeller O.D. is the correct distancefrom the cut water (Gap “B”).Corrections can be achieved by twomethods:

1.Overfiling: This disturbancemay be partly or entirely eliminatedby tapering the vanes by “overfiling”or removal of metal on the leading

face of the vanes (Figure 4). Thistechnique has the additional advan-tage of restoring the vane exit angleto near that of the maximum impellerdesign (i.e., before the diameter wasreduced).

2. Underfiling: Sharpening theunderside of the trailing edge of thevane (Figure 5) can enlarge the outletarea of the liquid channel. This willgenerally result in about five percentmore head near the best efficiencypoint, depending on the outlet vaneangle. At least 1/8 inch of vane tipthickness must be left. Sharpeningthe vanes also improves the efficien-cy slightly. Where there are highstage pressures, you must sharpenthe vanes carefully because the vanesare under high static and dynamicstresses. ■

Ed Nelson is a consultant to theturbomachinery and rotating equipmentindustries. He serves on the Pumps andSystems Editorial Advisory Board.

FIGURE 5. SHARPENING OF IMPELLER VANES Q:

A:oblique cut leaves the shroudsunchanged and solves the structuralstrength problem as well as improv-ing the exit flow pattern.

Normal sharpening

Leave at least 1/8 ”

Mill or grind away

Original outlet widthNew outlet width

Max. sharpening

Originalthickness

Page 42: Centrifugal Pumps Handbook

42 The Pump Handbook Series

cooling water application. Thecentrifugal pump was rated at1,000 rpm for a suction lift of 15ft against an 80 ft total headwhen running at 750 rpm. Usingthe rules governing the relationof capacity, head, and speed, wethought it should be possible toobtain 1,400 gpm against a totalhead of 150 ft by replacing theoriginal motor with a larger,1,050 rpm motor. However, inits new service, the pump hasnot provided anywhere near theanticipated capacity. What couldbe wrong?

Switching a pump fromone service to anotherfrequently appears tobe an easy and cost-effective way of avoid-

ing the purchase of a new pumpdesigned for the desired use.Unfortunately, such switches canbe tricky business (as discussed inthis column in March, 1993). Yetsince almost everyone will betempted to engineer such a switchat least once during a career, itmight be helpful to review keycalculations that are needed in aneffort to determine what wentwrong in this case.

Summarizing the informationyou provided, we have one pumpintended to handle streams ofcomparable quality (basically coldwater) that has been operatedwith two different motors.Knowing the original designcapacity and total head, we canquickly determine the same infor-mation for the new application bythe following equations:

Q2 = Q1 x (N2/N1)

and

H2 = H1 x (N2/N1)2

We have a 1,200 gpmcentrifugal pump thattransfers water from apublic reservoir up toour makeup reservoir

inside the plant. The change inelevation is about 30 ft over adistance of roughly a mile. Thepump does not operate continu-ously, rather it is turned on andoff by plant staff who check themakeup reservoir level once pershift. According to operators, thepump seems to deliver ”fullflow“ when first started, but isoperating at a much lower capac-ity when checked later. Whatcould cause this consistentdecline in pump capacity?

Your problem couldhave a couple of caus-es. One cause mightbe air that has enteredthe system and accu-

mulated in the pump. While it ispossible for the reservoir water tobe saturated with air that willcome out of solution in the pump,most centrifugal pumps can han-dle a small amount of air (2–3%by volume), which will passthrough as bubbles with the liq-uid. Instead, the introduction ofair is more likely equipment relat-ed. Depending on the pump andpiping system, air can get into thewater in several ways.

For the pump, check the shaftsleeves to ensure that the sealbetween the sleeves and theimpeller hub is adequate. Thenexamine the stuffing boxes. Forpumps operating with a suctionlift, lantern rings should beinstalled and have seal waterunder positive pressure.

Piping can be more difficult toexamine because most of it willlikely be underground. However,any surface piping should beinspected to assure that it is air-

tight. And the as-built drawingsshould also be studied to determine ifthere might be any irregularities(such as improper pitch or high spots)along the pipeline in which air pock-ets could form.

If air is the cause of reducedpump capacity, this can be confirmedby stopping the pump, opening andclosing the vent valve on top of thevolute, and immediately restartingthe pump (which should run at fullcapacity). Do not open the vent valvewhile the pump is operating. Even ifair is present in the pump duringoperation, it will be trapped near thecenter of the impeller while the heav-ier water will be forced to the outeredge (and out the vent valve if it isopen) (see Ask the Experts,November 1993).

A second possible source of yourdifficulty is the intake at the publicreservoir. From the information pre-sented, the system probably has asubmerged offshore intake with someform of screening to prevent theentrainment of unwanted materials.Underwater plants, particularly fila-mentous grasses, can be drawn intoand entangled in the intake screening,blocking flow. When the pump is notoperating, the natural underwatercurrents can clear some or all of theblockage so that full flow is temporar-ily restored at pump startup.

The intake can also contributeair to the system. If the level of thepublic reservoir has dropped, the dis-tance between the surface of thewater and the submerged intakemight not be adequate to prevent theformation of vortices whenever thepump is operating. Such vortices canbring significant amounts of air intothe system.

We recently decided tomove a relatively old butinfrequently used stand-by service water systempump to an auxiliary

Examining Pump Capacity ProblemsBY WAYNE C. MICHELETTI

Q:

Q:

A:A:

CENTRIFUGAL PUMPS

HANDBOOK

Page 43: Centrifugal Pumps Handbook

The Pump Handbook Series 43

to respond quickly and efficientlyto lower (and higher) flowdemands, enabling you to con-serve energy as well as water.However, electrical adjustable-speed drives can be expensive andshould be thoroughly evaluatedfrom an economic perspective for”older“ pumps.

A second, more commonapproach is simply to throttle thedischarge. Doing so will introducea new artificial friction loss com-ponent to the head. This will shiftthe present system–head curveupward to intersect the pumphead–capacity curve at a newoperating point (corresponding toreduced capacity). It should alsoreduce the energy requirementslightly.

It is never advisable to throt-tle the pump suction. Thisapproach (occasionally referredto as operating in the ”break“)changes the pump head–capacitycurve through cavitation. Theresulting operation is not onlyinefficient but potentially damag-ing to internal pump compo-nents.

The third option is moreenergy-efficient than throttling,but only suitable if a permanentreduction in pump capacity isacceptable. The pump impellercan be cut down, essentially low-ering the pump head–capacitycurve. However, before trim-ming an impeller, a number ofother factors and resulting impli-cations should be carefully con-sidered. (These were discussed inthis column in the January, 1993issue.) ■

Wayne Micheletti is a water andwastewater consultant and a memberof the Pumps and Systems EditorialAdvisory Board.

Using this equation and the origi-nal pump design data (Q = 1,000gpm; N = 750 rpm), the value of S is2,737. Since the suction specificspeed is constant for a given pump,this equation can be rearranged tocalculate the NPSHR for the pump’snew application (Q = 1,400 gpm; N= 1,050 rpm). The increase in pumpcapacity and speed mean an increasein the NPSHR from 17.8 ft to 34.9 ft.

As a result, the conditions of thenew application correspond to a suc-tion head as opposed to a suction lift:

barometric pressure (abs.) = 33.9 ft

– vapor pressure of water = 1.1 ft

+ suction lift = 2.1 ft

NPSHR = 34.9 ft

If the NPSH available (the differ-ence between the absolute suctionpressure and the liquid vapor pres-sure) is reduced below the NPSHrequired, then the pump capacity isreduced, and the pump is likely tocavitate. Unless you can change thesuction conditions for the auxiliarycooling water application, it would bebetter to buy a new pump thanattempt this switch.

As the result of a recent-ly implemented watermanagement program,several of our older, con-stant-speed centrifugal

pumps now provide significantlymore water than is required. Whatis the best approach for operatingthese pumps at reduced capacity?

Congratulations. Manywould envy the problemproduced by your successin water conservation.Fortunately, there are

three well-proven solutions to reduc-ing existing pump capacity.

If the system flow is expected tochange frequently or irregularly, newadjustable-speed drives might be inorder. For variable-torque applica-tions (such as centrifugal pumps),solid-state AC or DC drives are usu-ally best. They will allow the pump

Q:

A:

where

Q = pump capacity (gpm)

N = motor speed (rpm)

H = total head (ft)

As you expected, the corre-sponding capacity and total headfor the new motor (at 1,050 rpm)should be:

Q2 = 1,000 x (1,050/750) = 1400 gpm

H2 = 80 x (1050/750)2 = 157 ft

So far so good. According tothese calculations, the pumpshould be able to provide thedesired flow against the estimatedhead. But before a pump cantransfer any fluid, the liquid musthave enough outside energy toenter the pumping element at thevelocity corresponding to therequired pump flow rate. For acentrifugal pump, this energymust be great enough to make thefluid flow into the impeller eyewith sufficient force to prevent thefluid pressure from droppingbelow its vapor pressure whenpassing the inlet vane edge.

This outside energy require-ment is known as the Net PositiveSuction Head Required or NPSHR.Assuming that your system is atsea level, this value (for the origi-nal pump design) can be deter-mined as follows:

barometric pressure (abs.) = 33.9 ft

– vapor pressure of water = 1.1 ft

– suction lift = 15.0 ft

NPSHR = 17.8 ft

For centrifugal pumps, theNPSHR can be correlated to pumpcapacity and motor speed by avalue known as the suction specif-ic speed (S) according to the fol-lowing formula:

S = (N x Q0.5)/(NPSHR)0.75

Page 44: Centrifugal Pumps Handbook

44 The Pump Handbook Series

umps sometimes sufferdamage unnecessarilybecause they are not 100%full of liquid before they

are started. The systems in whichthey function either are not or can-not be completely vented. A com-mon misconception is that a pumpthat produces discharge pressureimmediately after start-up was suf-ficiently full of liquid. For someusers, this is the working definitionof the word “primed.”

Igor J. Karassik, an interna-tionally recognized authority onpump systems, has written foryears about the need to remove allof the gas or vapor from pumpsbefore starting them.

Widespread understanding ofthe problems trapped gasses cancause developed during 1991 froman effort to understand erosionproblems with enlarged, tapered-bore seal chambers used on ANSIB73.1M chemical pumps. Testing,independently performed by and

for a number of pump and seal com-panies, showed that the liquid in theseal chamber circulates around thechamber at a large fraction of shaftrotation speed. A secondary flow wasobserved heading away from theimpeller, along the outside diameterof the seal chamber, and toward theimpeller along the shaft. Together,these two flow patterns explainedhow erosion damage was occurringin a few cases where abrasive solidswere present (Figure 1).

An unexpected byproduct of thistesting was the realization that gas orvapor that is present in the sealchamber at the time the pump isstarted can be trapped there for sev-eral minutes by these same flow pat-terns. Worse, trapped vapors or gasestend to accumulate close to the shaft,near the rear of the seal chamber.For most single mechanical sealinstallations, that is where the sealfaces are located.

Here are some common ques-tions on venting pump systems:

Why would gas endup close to the shaft?

Imagine the seal cham-ber is more than halffull of liquid as thepump is started. As thepump shaft (and

mechanical seal) picks up speed,viscous drag causes the liquid tobegin to circulate around thechamber. Soon, centrifugal forceovercomes gravity, and the liquidis thrown to the outside of the sealchamber. Any gas is forced inwardby the denser liquid.

What problems arecaused by the gas?

The most commonproblem is mechanicalseal damage. If the gasbubble is big enough tosurround the seal faces,

it can prevent the liquid in the sealchamber from cooling and lubri-cating the faces. Large pockets ofgas can damage wear rings andbushings, but gas would tend to beswept out of these areas quickly.

My pumps have flood-ed suctions. Won’tthey fill completelywhen I open the suc-tion valve?

Probably not! While it istrue most most modernpumps are designed tobe completely self-vent-ing, there is an assump-

tion that there is someplace for thegas to go as the liquid enters.Unless the discharge valve isopened slightly and there is no dis-charge pressure, the gas hasnowhere to go. When a horizontalend-suction pump is installed (orre-installed after a repair), and the

Venting Pump SystemsBY MICHAEL D. SMITH

P Q:A:

Q:A:

Q:A:

FIGURE 1

Primary and secondary flow patterns can result in erosion damage.

CENTRIFUGAL PUMPS

HANDBOOK

Page 45: Centrifugal Pumps Handbook

The Pump Handbook Series 45

suction valve is opened, it willoften fill to the top of the suctionpipe. When the gas (air, in thiscase) can no longer escape out thesuction pipe, it will compress asmall amount in response to thesuction pressure. A very large gaspocket remains in the pump atthis point, although the pump isprobably “primed.”

Why has this cause ofseal damage remainedhidden?

A big reason why pock-ets of gas have not beena concern is that theydon’t always cause animmediate failure. Seal

face damage progresses each timethe pump is started while it is notfull. Venting is not an issue inmany pump starts because thepump was not drained since itslast use. If a pump seal fails aboutonce a year, we assume it has aone-year wear life. We don’t evenconsider that it might be failingevery third time the pump is start-ed without being 100% full of liq-uid.

How can the problembe avoided?

The operator must under-stand why it is importantto fill the pump complete-ly. The pump system must

Venting Your Pumps

It is common to have ventingproblems when a pump is con-nected to a system that is pres-surized even when the pump isnot running. These systems oftenemploy a check valve in additionto a discharge valve. Some usersdrill a small hole in the checkvalve flapper to help vent thepump, but this technique is noteffective for the most commonoperating strategies.

The following is a simpleprocedure that can be used to getmore complete venting of thesehard-to-vent systems. It assumesthe pump is empty of liquid andboth suction and discharge valvesare closed.• Open suction valve (pump

fills part way).• Close suction valve.• Open discharge valve part

way (once pressure equal-izes, air will rise into dis-charge piping).

• Open suction valve.• Start pump.

CAUTION:The pump seal will be

exposed to full discharge pres-sure using this procedure.

Never start a pump with thesuction valve closed.

be capable of being completely vented.When the liquid can be released to theatmosphere, a vent valve is all that maybe required. See the sidebar at the endof this article for a procedural solution toa common situation. While discussingsystem design, it should be noted thatthe suction line should not have anyhigh points. The suction line should risecontinuously either toward the pump orback to the source. If a local high spot isnecessary, it will also have to be vented.I have seen many long suction lines thatwere designed to be level that still hadlocal high spots several pipe diametersabove the ends. This can be due to prob-lems with the original installation or theshifting of pipe supports at a later time.

CONCLUSIONWhoever has responsibility for

the design of the “system” will needinformation on the pump, the piping,and the operating conditions to assurethat it can be vented. ■

Michael D. Smith is a SeniorConsulting Engineer at the DuPontCompany in Wilmington, DE.

Q:A:

Q:A:

Page 46: Centrifugal Pumps Handbook

46 The Pump Handbook Series

BY JOHN W. DUFOUR AND LYNN C. FULTON

Installation and Start-Up Troubleshooting

lot of time and money arespent manufacturing and test-ing centrifugal pumps anddeveloping purchasing specifi-

cations for bidding and selectingthem. However, events after leavingthe manufacturer may result in apump that won’t perform reliably ordeliver the desired hydraulics.

SHIPPING AND HANDLING

Once the pump/driver/baseplateassembly leaves the factory, anythingcan happen if specific instructions onhow it should be shipped, received,stored, and installed are not followed.A document that records what wasand must be done, what must beapproved and by whom, and whenthese events should happen is crucial.Without this, work will be missed orduplicated.

Manufacturers prepare productsfor shipping differently. Some mountpumps in custom-made crates, whileothers hang the shipping tag on auxil-iary piping and bolt two-by-fours tothe base. The purchaser should definespecial requirements. Will the pumpbe shipped overseas? Is long-term stor-age required? Is there lifting equip-ment at the site? These questions mustbe answered ahead of time.

In all cases, Material Safety DataSheets should be included duringshipping and installation. Everyonewho comes in contact with the pumpneeds to know what’s in it.

There are other questions, too.What form of transportation will beused? A dedicated truck or a com-mon carrier? Who will receive theequipment? When? A dedicatedtruck usually has two drivers dri-ving around the clock, directly frommanufacturer to delivery site. Thisis costly but quick. A common car-rier is less expensive but can takelonger. For example, pumps froman East Coast manufacturer, des-tined for Texas, were loaded on atruck Friday afternoon. The pumpsarrived 15 days later. With no oneto receive them, the driver leftthem at a warehouse. It took two

days to locate them, and they weredelivered a week after that.

Pumps are easily damaged dur-ing transportation, storage, orinstallation. Most baseplates aredesigned to be lifted with an over-head device or moved by fork lift.Care must be taken to prevent dam-age to auxiliary piping from liftingslings or hooks. Storage facilitiesoften don’t have an overhead crane,so a forklift moves the assembly offthe truck and around the storagearea. Again care must be taken tobalance the load before lifting andto avoid bumping or dropping theassembly (falling just an inch cancrack the mechanical seal facering). Never lift the pump by itsshaft or auxiliary piping.

STORAGE

Sometimes the pump goes direct-ly from truck to foundation, but theassembly is often stored for a time.Storage may be a graveled yard or awarehouse with overhead liftingequipment and a controlled environ-ment. In any case, following threerules will help avoid problems:

1. Keep oil/grease in the bearings.

2. Keep water/moisture out of thecase (seal, windings, etc.).

3. Protect the pump from abuse.

Check the pump over. To pre-vent baseplate distortion, place it leveland out of traffic. See that all coverplates are bolted on. Be sure no auxil-iary piping or components were lostor damaged in transit; replacing a partmay delay start-up. Bearing housingsshould be filled with oil to the bottomof the shaft and rotated periodically tokeep bearings coated. Document whoturns it and when. Pumps stored long-term with oil mist lubrication shouldbe hooked to a portable mist genera-tor. Verify that the mechanical sealsleeve locking collar is tight and thatthe shaft turns freely.

Stored drivers may require extracare. Heaters on electric motors shouldbe energized to keep windings dry.

Steam turbines often have carbon ringsand seals. Remove them to preventcorrosion under the rings, or continu-ously purge the case with dry nitrogen.

INSTALLATION

As mentioned, prepare a docu-ment to ensure proper installation.Outline specific requirements, insequence, for each pump. Definetasks and inspections, who is respon-sible, and special procedures—grout-ing plans, cold alignment targets,pre-start-up checks, hot alignmentchecks, etc.

Vendors often give details oninstallation, and writings on the sub-ject are available. Here is a list to aidinstallation:

GROUTING

• Prepare the foundation surface.Chip latence off, exposing aggre-gate. Remove loose material,grease, and water.

• Level the baseplate using jack-bolts bearing on jackplates(Photos 1 and 2). Jackplatesshould have rounded corners.It’s easier to slice sections fromround stock than to cut plate.

• Remove pump and motor beforeinstallation; it’s easier to level thebaseplate and pour grout.

• Check the baseplate bottom forcleanliness. Verify that eachcompartment has grout andvent holes. Drill holes beforelifting the baseplate onto thefoundation.

• Don’t grout around anchor bolts.Baseplates are grouted to provideuniform load distribution. Anchorbolts hold the pump down. Tokeep anchor bolts free to stretch,install sleeves around bolts.

• Install the baseplate, establishingcorrect elevation (within 1/8 in.)and pedestal level (within 0.002in./ft). Some contractors like toput pumps back on the baseplateto shoot the nozzle elevation.

A

CENTRIFUGAL PUMPS

HANDBOOK

Page 47: Centrifugal Pumps Handbook

The Pump Handbook Series 47

This is unnecessary and may dis-tort the baseplate.

• Coat forms with furniture pastewax to ease removal. Fix formsto the foundation block at differ-ent elevations to avoid fracturelines from anchor studs. Drilledholes with screws look betterafter removing forms and elimi-nate potential impact cracksfrom hammering nails or usingcharged drivers.

• Tape or grease machined mountingsurfaces for protection.

• Ensure that grout flows into allcompartments by using a headbox and vent tubes. The headbox can be six-inch sonotubesRTV’d to the baseplate surround-ing the pour holes. Vents can beplastic pipe. These should be atleast six inches high to provideenough head to get all voidsunder the baseplate.

• Grout between 60 and 90°F(Photo 3). Cooler temperaturesdon’t allow curing. Higher tem-peratures may cause fast curingand heat cracks. Grout shouldharden in 24 hours.

• As soon as the grout firms (nothardens), remove vent pipesand head boxes. Grout consis-tency should be like hard rub-ber, making it easy to trim.

• Forms can usually be removedafter 48 hours. Remove jack-bolts from baseplate and fillholes with RTV.

MOUNTING/ALIGNMENT• Set pump on its pedestals, center

bolts in their holes, and snug.This allows movement if side-to-side motor adjustment can’tachieve alignment.

• Mount motor with a minimumof 1/8 in. stainless shims underthe feet using the required dis-tance between shaft ends(DBSE). This is usually found onthe general arrangement or cou-pling drawing. With sleeve bear-ing motors, the magnetic centerof the motor with respect to thestator must be determined and

the DSBE set with the motorrotor in its magnetic center.

• Make sure the mechanical sealdrive collar locking screws aretight, then roll locating cams outof the drive collar. Lock cams outof the way or remove them.Remember that future work willrequire cams to reset seal com-pression—don’t loose them. Theshaft should turn freely.

• Align motor to pump, free ofany piping, using, as a mini-mum, the reverse indicatoralignment method. To avoidsoft foot, minimize shims undereach support area. When align-ment is achieved, tighten hold-down bolts and recheck.

• To check for soft foot, place adial indicator on each mountingfoot, then loosen the hold-downbolt. If the reading changesmore than .001 in., reshim.

PIPINGCare in fabricating and aligning

piping avoids problems that mayrequire recutting, fitting, rewelding,and retesting the pipe or lead to pre-mature pump failure. Good systemdesign supports piping loads andforces along spring hangers andbracing that don’t have to beremoved during normal mainte-nance. The system should be fabri-cated starting at the pump flanges,working toward the pipe rack, usingtemporary braces/supports to avoidpump strain.

The most common piping fabri-cation error, producing the largestpiping strain, is nonparallel flangefaces. A feeler gauge helps detect this.If you see a difference in two facingflange planes, piping strain willresult. For example, during installa-tion of circulating water pumps in arefinery, suction piping was forced tothe pump flange without checking fornon-parallel faces. The resultingstrain distorted the casing to the pointwhere the shaft and impeller wouldnot turn. Fortunately, no seriousdamage occurred. The cases werereclaimed after the piping wasaligned and supported.

• After fabrication and pipe test-ing, remove temporary bracingand lock-pins from spring hang-ers and check strain.

• Remove flange covers andinspect the pump for debris.Clean out the case. Bring thepiping to the pump flanges.Flange holes should dropthrough with no binding.

• Place dial indicators to monitorvertical and horizontal move-ment of pump shaft relative todriver shaft. Make up suctionand discharge flanges separately,continuously observing indicatorreadings. If movement exceeds0.001 in., piping strain is exces-sive. Readjust pipe, retighten,and retest.

PREOPERATIONAL CHECKS

The period from installation untilfull operation may be the mostimportant phase of pump life. It’sfilled with activity and riddled withpitfalls that can complicate start-upand prevent establishing a reliablesystem. The rules above also applyhere:

1. Keep bearings lubricated.

2. Keep moisture out of the case.

3. Protect the pump.

Drain and flush bearing housingswith clean oil. Oil rings may havemoved during handling, so lookthrough the vent caps to verify thatthey’re in position. On oil mist instal-lations see that mist reclassifiers havebeen installed correctly. Directed oilmist fittings have a “V” at the orifice.This must be pointed towards thebearing. Insure that all mist lines aresloped so no low points cause liquidbuildup and block flow.

Greased bearings should berepacked with the correct grease.Make sure all old grease is displacedby new. Different greases (lithium-vs. soda-based) have incompatibleadditives. Mixing two greases cangive an inferior blend.

Bump check motors for properrotation. Do not attempt this whilethe motor is coupled to the pump.Reverse rotation can cause theimpeller to loosen or come off the

Page 48: Centrifugal Pumps Handbook

48 The Pump Handbook Series

shaft. If rotation is correct, run themotor alone for at least one half hour.Monitor bearing temperatures andmotor bearing housing vibration. Thisshould reveal any major problems.Most others will not be revealed untilit is loaded and generating heat.Install the coupling spacer and guardand verify smooth assembly rotation.

Don’t overlook small steam tur-bine drivers. Verify rota-tion direction by inspectingnozzle orientation. As soonas steam is available,check the operation of thegovernor and overspeedsystems. Run the turbinesolo at least one half hour.

Mechanical seals andbearings are easilydestroyed during initialstart-up on hot pumpswhere water is circulatedfor cooling. Install pressuregauges, temperature indi-cators, and valves so waterflow can be regulated andadjusted. Throttling valvesare typically installed onparallel outlet lines toadjust flow to each pumpskid area. It’s a good ideato flow most cooling waterthrough the seal and bear-ing coolers initially.

Debris in the pumpand seals is a problem dur-ing initial start-up. A welding rodlodged in an impeller eye can seizethe pump. To prevent this, use suc-tion screens. Insert temporary stain-ers if they’re not built in. Pressuregauges on both sides of the strainerindicate when it is plugging.

INITIAL OPERATIONThe electronic equipment adage,

“If it works the first hundred hours itshould work a lifetime,” also appliesto centrifugal pumps. Knowledge ofthe equipment and the system it willoperate in are key to successful start-up. While each installation is differ-ent, this general procedure will helpprevent problems:• Close case drains and vents.

Slowly open the suction line.Look for leaks at the caseflanges, seal area, seal piping,drain piping, and inlet dis-

charge piping. If there areleaks, return to a safe situationand repair them. If leaks occuraround the shaft, determine ifseal faces are leaking or if theleak is under the seal sleeve.Stop leaks between sleeve andshaft by adjusting the drive col-lar. Stop leaks around the sealflange by retorquing the bolting

to clamp the sta-tionary gasketing.It’s important toknow where theleak is beforepulling the pumpapart.

• If no leaks areseen, open theinlet valve 100%.Vent areas of thesystem that don’tself-vent. Crackthe dischargevalve. Start-uphorsepower isminimum to theleft hand side ofthe pump curve.On systems pump-ing higher specificgravity liquids atstart-up than dur-ing normal opera-tion (typical ofcold start-ups), the

discharge system may have tobe throttled to avoid motoroverloading. Throttling to 50%BEP is acceptable in most cases,but more than that may causeseal problems.

• Start the pump. Slowly openthe discharge valve. If a dis-charge control valve is installedand on automatic, the controlvalve will be wide open untilthe block valve opens enoughfor the control system to takeover. If the pump cavitates,there may be too much flow.Start to pinch down on the dis-charge valving, preferentiallyusing the control valve. Mostsystems have a flow meter.Flow can sometimes be deter-mined from the meter directly,and the differential across thepump can be determined using

pressure gauges on the pump.Using flow and differentialhead, determine where thepump is operating on thecurve. Low flow, high headmay indicate running too farback, leading to bearing or sealproblems. High flow, low dif-ferential head means the pumpis running out on its curve andcould cavitate. Check differen-tial across the inlet screen anduse a spare pump before lowsuction pressure causes cavita-tion.

• Where flow can’t be measureddirectly with a meter, estimateit using motor current, horse-power requirement of thepump, and plotting that pointstraight up to the performancecurve. The intersection of thevertical line from the horsepow-er curve to the performancecurve should be the capacitypoint as long as specific gravityis similar to horsepower curvespecific gravity. Differentialhead on the pump should besimilar to differential head onthe curve at the capacity pointdetermined from the horsepowercalculation.

From Figure 1:

Example point 1.

P2 – P1TDH = 2.31 —————— = 188 ft

S.G.

where

P1 = 3 psig

P2 = 60 psig

S.G. = 0.7

√ 3 x I x V x ηM.H.O. = ——————— = 26.5 Hp

746

where

M.H.O. = motor horsepower out-put

V = motor voltage = 460 V (30 Hpmotor, 3φ, 460 V [line toline])

Good systemdesign

supportspiping loadsand forces

along springhangers andbracing thatdon’t have tobe removed

during normalmaintenance.

Page 49: Centrifugal Pumps Handbook

The Pump Handbook Series 49

η = motor efficiency = 90% (2-pole motor, 90% efficiency)

I = phase amp measurement =27.2 amps

Example point 2.

Throttling pump discharge

P2 – P1TDH = 2.31 —————— = 214.5 ft

S.G.

whereP1 = 3 psig

P2 = 68 psig

I = 20.8 amps

√ 3 x I x V x ηM.H.O. = ——————— = 20 Hp

746

where

M.H.O. = motor horsepower out-put

V = motor voltage = 460 V (30 Hpmotor, 3φ, 460 V [line toline])

η = motor efficiency = 90% (2-pole motor, 90% efficiency)

I = phase amp measurement =27.2 amps

• Check motor and pump vibra-tion. Vibration levels should bebelow 0.15 in./sec. Most newequipment vibrates less than 0.1in./sec. true peak.

• Compile documents for eachpump and file them for refer-ence. ■

John Dufour has more than 20years of experience working withmechanical equipment. He is ChiefEngineer, Mechanical EquipmentServices, for Amoco Oil Co, and isresponsible for specification, selection,installation, and consultation for rotat-ing equipment throughout the compa-ny’s refinery, pipeline, and marketingoperations. He holds bachelor’s degreesin metallurgical engineering and engi-neering administration from MichiganTechnological University. Mr. Dufouralso serves on the Pumps and SystemsEditorial Advisory Board.

Lynn Fulton is a professionalengineer registered in Indiana andIllinois. He has been with WhitingEngineering more than ten years, inmechanical services and maintenance.He has a bachelor’s degree inmechanical engineering from theUniversity of Illinois at Chicago andis chairman of the Chicago chapter ofthe Vibration Institute.

Point 2

Point 1

250

200

150

100

50

0

Tot

al H

ead

(ft.)

0 100 200 300 400 500 600Flow (gpm)

40

30

20

10

0

BH

P

FIGURE 1

Head and BHP vs. Flow. Operating point 1: using the BHP vs. flowcurve with horsepower calculation derived from measurement ofcurrent with voltage assumed to be 460 V, flow is found to be 405gpm. The calculated 188 ft based on pressure differential confirmsflow to be 405 gpm. Operating point 2: similar calculations forhorsepower and head at operating point 2 also confirm the calcula-tion method. See text under “Initial Operation” for calculations.

Page 50: Centrifugal Pumps Handbook

50 The Pump Handbook Series

FIGURE 1pgrade (up + grade, v.): toraise the grade of; to raise thequality of a manufactured prod-uct (Webster’s Third New

International Dictionary).A pump upgrade (also called a

revamp or retrofit) involves changingmechanical or hydraulic design ormaterials to solve a problem or increasereliable run time. An upgrade is differ-ent than a repair, which attempts toduplicate original construction anddesign, whereas an upgrade improvesthe design beyond the original.

Rerates are a type of hydraulicupgrade, usually involving a changein pump head capacity. Repoweringmay involve repairs and/or upgrades.Philosophically, repowering is differ-ent from normal pump maintenancebecause the plant being repoweredhas decided to spend capital moniesto extend the plant’s useful life. Plantsbeing repowered are candidates forpump upgrades because they areexpected to run reliably with highcapacity factors and can justify theadditional cost (above and beyondnormal repairs) to upgrade pumps.

Pump upgrade goals include:• decreasing plant operations and

maintenance expenses• increasing mean time between

failures (MTBF)• increasing pump and plant avail-

ability• increasing pump efficiency• complying with the latest legisla-

tive mandates (such as the Clean

Air Act Amendments of 1990)• minimizing the risk of fire or other

safety hazards• eliminating hazardous materials

Pump upgrades can be dividedinto major categories:• mechanical design• hydraulic design• material• ancillary/system

THE UPGRADE PROCESS

To identify upgrade candidates,pump users should review mainte-nance records to see which pumpswere responsible for a disproportion-ate share of expenses or caused safe-ty or reliability concerns. Once thesepumps are identified, work with theupgrade supplier to identify and eval-uate upgrades available for your par-ticular pump. Provide the supplier

BY KURT SCHUMANN

U

Improvements to make your equipment better than new.

Upgrading Utility andProcess Pumps

Original residual heat removal pumps in safety service atnuclear power plants. The design has high maintenancehours and exposure dosage due to short mechanical seal lifeand an overhung shaft design that makes seal maintenancedifficult.

CENTRIFUGAL PUMPS

HANDBOOK

Page 51: Centrifugal Pumps Handbook

The Pump Handbook Series 51

- replace nickel-aluminum-bronzeparts with austenitic stainless steel- use other special alloys for criti-cal parts- install non-metallic bearings

Ancillary/systems upgrades:• install vibration monitoring and

recording instrumentation• improve lube oil system and

instrumentation• modify seal injection• perform pump intake scale model

testingThe following are examples of

upgrades to improve pump operation:

RHR PUMP COUPLING MODIFICATION

These pumps are residual heatremoval pumps, close coupled designin safety service at nuclear powerplants (Figure 1). The original designresulted in considerable time,expense, and man-rem exposure fornormal maintenance activities likeseal change-out and motor thrustbearing replacement. There was alsoa high risk of damaged equipmentfrom the difficulties of rigging incramped quarters. Pump upgrade kitsadd a bearing and a spacer coupling.Installation of these kits allow sealremoval without pump disassembly.

The original design frequentlyresulted in seal change-out timeslonger than the 72 hours permitted bymost plant safety evaluations. Theupgrade easily accommodates a sealor motor bearing change in 72 hours,without the high man-rem dosageinvolved in pump casing disassembly.This reduction in personnel exposureis an important benefit in any case,but it is especially so given theincreased industry focus on the issue.

These coupling modificationshave been supplied to several utilitycompanies as bolt-on hardware kitsinstalled during short outages (1 or 2weeks). Other items like oil drain loca-tion and mechanical seal venting havealso been improved in the design.

BOILER FEED PUMPS

Boiler feed pumps are at theheart of most power plants, and eco-nomical plant operation depends onreliable pump operation. Manypumps from the utility building boomof the 1950s–70s had larger capacities

FIGURE 2

An upgrade of Figure 1. Spacer coupling between pumpand motor allows seal access without disassembling thepump. A bearing above the seal limits shaft deflection.Conversion from non-cartridge mechanical seal to cartridge-type seal eases assembly. Benefits: increased seal life,decreased leakage, and decreased personnel exposure whilechanging seals.

with a maintenance history so prob-lem areas can be addressed.

Plan for a future outage wherethe plant or process will be shutdown long enough to completedesign and hardware changes. This isimportant—upgrades take time toengineer and implement, and theymust be planned in advance based onrepair or outage schedules. Thisapproach can also save money;upgrading worn out parts during thenormal repair cycle (instead of replac-ing components that still have lifeleft) minimizes incremental cost.

The following are upgrade exam-ples from each of the areas above:

Mechanical design upgrades:• install a stiffer shaft/rotor to

reduce vibration• modify structural elements to

remove natural frequencies fromthe range of pump forcing fre-quencies (rotational frequency,blade pass frequency, etc.)

• eliminate threads (a source ofbreakage) on pump shafts

• modify components to makeassembly/disassembly easier

• convert mechanical seals to fur-ther restrict or eliminate leakage

Hydraulic design upgrades:• redesign first stage impellers to

reduce cavitation damage• redesign impellers to lower vibra-

tion for part load/peaking opera-tion

• control “A” and “B” gaps to reducepressure pulsations and vibration

• improve efficiency• optimize blade number to reduce

pressure pulsations and vibration• increase pump head capacity to

meet system requirements

Material upgrades:• eliminate asbestos, an environ-

mental hazard• install impellers made of cavita-

tion-resistant materials for longerlife

• use hardened wear parts toincrease MTBF

• eliminate leaded bronzes becauseof environmental problems withlead

• improve product-lubricated bear-ing materials

• change materials for seawater use- replace Monel components withaustenitic stainless steel

Page 52: Centrifugal Pumps Handbook

The Pump Handbook Series 53

overlap can be modified to cur-rent standards. By evaluatingthe cost of these modificationsin light of expected benefits,users can choose the modifica-tion required to meet processrequirements with the least cost.

6. Improved first-stage impellerinlet designs expand the stableoperating range to lower flowrates without cavitation dam-age, vibration, or pressure pul-sations. Material upgrades forfirst stage impeller service resistcavitation damage while main-taining ductility, corrosion resis-tance, and weld repairability.

7. Dry couplings (like the flexibledisc or diaphragm-type) elimi-nate the need for periodic lubri-cation and the associated chanceof failure. They are often lighterthan gear couplings, resulting inlower vibration, and they toleratemore misalignment than gear-type couplings.

8. Instrumentation can be added tomonitor and protect pumps.Possibilities range from simplevibration/temperature switchesto complete monitoring of all keyoperation variables, includingremote monitoring, diagnostics,etc. The most common itemsmonitored include:a. shaft or bearing cap vibrationb. lube oil temperature and pressurec. axial and radial bearing temperatured. casing (or barrel) temperature

Additional items include:e. pump suction condition (pressure, temperature)f. pump discharge condition(pressure, temperature)g. pump flow rateh. horsepower, efficiencyi. balance drum leak off (tem-perature, flow rate)j. seal (drain temperature,mechanical seal face temperature,stuffing box temperature, etc.)

Monitoring can be stand-alone or canfeed into the plant’s control system.9. Some older pumps have open

vane diffusors. Vanes can fatigue

due to unsteady hydraulic loads.Using a shrouded diffusor elimi-nates breakage problems andallows improved alignment.

10. Improved bearing designs are avail-able, including ”high stability“designs to eliminate half frequency(”oil whirl“) problems. Special atten-tion is paid to individual plant oper-ating modes (low-speed operation,turning gear, etc.) in recommendinga particular bearing design.

CIRCULATING WATER PUMPSCirculating water pump mainte-

nance requirements vary greatly,depending on whether the pumps areused in freshwater or seawater.

For most freshwater applica-tions, typical problems requiringpump maintenance are excessivevibration and premature bearingwear.

Vibrations can be analyzed usingmodal analysis or standard spectrumanalysis techniques to identify theroot cause of the vibration. If neces-sary, a finite element analysis (FEA)model of the pump can be built andcorrelated to the field data to verifythe cause. It can also be used to helpredesign the pump.

To improve bearing and sleevelife, upgrades are available to increasewear resistance through material selec-tion and hardcoating.

Circulating water pumps in sea-water face additional problems due tocorrosion. Material selection is criti-cal, and the selection process mustconsider general corrosion as well asvelocity effects, galvanic compatibili-ty, and pitting resistance, plus manu-facturability and cost.

The cost difference betweenmaterials can be significant becausethese pumps are large; care must betaken not to over-specify materialsand inflate the price of equipment formarginal benefits. In some cases,lower-cost material may be more reli-able. For example, 316 stainless steelhas better pitting resistance thanMonel in seawater, yet Monel ismore expensive.

Upgrades for circulating waterpumps include (Figure 4):

1. An inner column stop on pull-out style pumps to hold downthe pump element during start-

ing, stopping, unit trips, andother transients.

2. A flanged inner column withrabbet fits replaces screwedinner columns, resulting in bet-ter bearing alignment and easierdisassembly.

FIGURE 4

Circulating water pump upgrades.See text under ”Circulating WaterPumps“ for details.

1

2

3

4

5

6

8

9

7

Page 53: Centrifugal Pumps Handbook

54 The Pump Handbook Series

3. Bearing spiders provide stifferbearing support.

4. Shroud metallurgy upgraded inhigh-velocity areas eliminateserosion and corrosion damageand extends efficient pump life.

5. Inlet bell modifications lower therequired submergence and reducevortexing. Intake studies can beperformed to correct vortexingand other inlet problems and giveuniform flows to the pump,resulting in stable operation.

6. Modified impellers optimize cool-ing water flow, increase plantoutput, or save pumping horse-power. Upgraded impeller materi-als resist erosion, corrosion, andcavitation damage.

7. Erosion in iron casing vanes canbe repaired. Coatings can beapplied to extend casing life.

8. Keyed shaft coupling improvesshaft alignment and eliminatesproblems associated withremoving threaded couplings.Shafts can usually be re-machined and re-used.

9. Rabbet-fit drive couplings replacebody-bound bolt couplings andimprove alignment repeatability.

BOILER CIRCULATING PUMPSBoiler circulating pumps (BCPs)

are in a particularly severe duty,handling 600°F water at over 3,000psig.

Many of the original pumps sup-plied in the 1950s and 60s exhibitedless-than-desirable life spans. In lightof this, an upgrade program wasdeveloped that:• adds a graphite-impregnated

bearing• improves the primary sealing

device• incorporates other reliability

”lessons learned“ (Figure 5)To date, more than 200 BCPs

have been upgraded to this newdesign, and the resulting MTBF istypically two to three times that ofthe pumps before upgrading.

API PROCESS PUMPSProcess pumps may handle

hazardous materials, and as a resultseal leakage is critical. Industrystandards (API 610 7th Edition), aswell as federal legislation like the

Clean Air Act, address mechanicalseal reliability and pump mainte-nance.

There are up-grades available forprocess pumps toimprove seal reliabili-ty and reduce emis-sions (Figure 6),including:• using a heavy-

duty rotor (shaftand bearings)

• enlarging thestuffing box borefor better sealcooling

• using a heavy-duty cartridge seal(single, double, ortandem)

Another option, if extremelylow levels of emissions are required(for instance, pumping benzene, acarcinogen), is to use sealless (mag-netic drive) technology. This can beaccomplished by repowering (re-using the casing, bedplate, and dri-ver, along with upgraded pumpinternals) or replacing the wholepump.

SUMMARYThese descriptions cover some

typical upgrades. This article focusedon specific types of pumps, butupgrades are available for most modelsand sizes. Pump companies are usefulresources for aid in problem solving.They are usually anxious to apply newtechnology and gain field experiencewith new designs and materials. Mostupgrade suppliers can customizeupgrades for individual users. Reviewyour maintenance problems and dis-cuss them with your pump supplier.

Pump upgrades are a cost-effec-tive way to improve plant perfor-mance within budget constraints.When upgrades are properly per-formed, an upgraded pump may wellbe ”better than new.“ ■

Kurt Schumann is Manager ofPump Upgrades for the EngineeredPump Group of the Ingersoll-DresserPump Company, located in Phillipsburg,NJ. He has 18 years of experience indesign engineering and field service ofutility and process pumps.

FIGURE 5

Right: Original boiler circulatingpump design. The throttle bushingand throttle sleeve wear quickly,reducing floating seal ring life and shortening service intervals betweenpump rebuilds. Left: The upgradeddesign. The retrofit incorporates awater-lubricated carbon bearing andredesigned floating seals.

FIGURE 6

Pump modification kits upgrade API 5th and 6th edi-tion process pumps with the features of the 7th edition

New Design Original Design

ThrottleBushing

Auxiliary FillConnection

Leak-Off toLow Pressure

Heater(Deaerator)

Injection fromdischarge ofboiler feed

pump

Reduced Shaft TIR Large Seal Chamber

Cartridge Seal Steel BearingHousing

Small SealChamber

Non-CartridgeSeal

Cast Iron BearingHousing

Existing A-Line (6th Edition)

A-Line 7th Edition Upgrades

Page 54: Centrifugal Pumps Handbook

The Pump Handbook Series 55

Variable Speed PumpingVariable speed pumping can save you money if you select and use systems wisely.

ost users operate their cen-trifugal pumps at a fixedspeed and accomplish anyrequired changes of flow

by using a throttling valve. This prac-tice is much like driving an automo-bile with the accelerator fullydepressed and changing speed bystepping on the brake!

There is a better way to drive anautomobile and there is a better wayto accomplish variable flow for a cen-trifugal pump. Variable speed motorsand associated electronic drives canbe used to adjust pump speed to pro-duce exactly the desired flow andhead. By varying the speed of thepump, users can enhance perfor-mance, save energy, eliminate theneed for throttling valves and reduceinputs of heat to the pumped liquid.

But to achieve these advantages,you must properly select the compo-nents of a variable speed system. Andproper selection requires a thoroughunderstanding of pump, motor anddriver designs for variable speedoperation.

BEHAVIOR OF VARIABLE SPEED PUMPS

A good place to begin a discussionof variable speed pumping is theinteraction between variable speedpumps and the fluid handling system.These interactions are different fromthose of a fixed speed pump.

For a fixed speed pump with flowcontrolled by a throttling valve,process demand depends on systemback pressure and piping resistance,as shown by a fixed system curve(Figure 1). Pump performance is also

represented by a fixed curve. Withthe discharge throttling valve fullyopened, the pump seeks equilibriumwith the system (point 1 in Figure 1:flow = Q1 and head = H1).

To change the flow to Q2, thethrottling valve is partially closed,changing the steepness of the systemcurve as seen at a point between thepump and the valve (at B-B in Figure 1). Closing the valve causesthe pump to “run back” on its curveto point 2, producing flow Q2 asdesired. The pump, which can onlyoperate on its fixed curve, produceshead H3 at point B-B. The pump thusproduces H3 at Q2 but only H2 at Q2

is delivered to the system. The addi-tional head (H3 - H2) is wasted acrossthe valve in the form of heat andnoise.

For a variable speed pump, flowis changed by varying speed. Thevariable speed pump retains itscharacteristic performance curveshape, changing flow and head inaccordance with the well-knownaffinity laws (Figure 2). With vary-ing speeds, pumps have widerangeability and thus any head-flow combination within the enve-lope can be achieved. And withappropriate precautions, pumpscan be operated at even higher orlower speeds than those shown onthe curve.

The shape of the system curveinfluences the amount that flow

will change with achange in speed. Flowis proportional to speedif no static lift existsbut not proportional tospeed if static lift exists(Figure 3). In systemswith static lift, a mini-mum speed existsbelow which the pumpwill produce no flow.

Such behavior doesnot violate the affinitylaws. It simply reflectsthe interaction of theshape of the systemcurve with those laws.In fact, it’s this interac-tion that makes variablespeed pumping advanta-geous (which also illus-trates that users must

understand these interactions).

Fixed speed centrifugal pump operation

FIGURE 1

BY STEPHEN MURPHY

M

CENTRIFUGAL PUMPS

HANDBOOK

Page 55: Centrifugal Pumps Handbook

56 The Pump Handbook Series

BENEFITS OF VARIABLE SPEED PUMPING

Because variable speed pumpscan produce a desired head andflow over a broad range ofhydraulic conditions, users do nothave to be as certain of requiredflow when they select a pump.Instead of finding the exact fixedspeed pump for the job, they caninstall a variable speed pump andadjust the speed to produce theexact conditions they require.

For example, one user requiredPump A to produce 125 gpm flow at2500 ft head in an upset condition and 100 gpm at 1500 ft under normalcon-ditions and Pump B for a 125 gpmflow at 1500 ft head under normalconditions. The user needed aninstalled spare for each pump, for atotal of four pumps. But by specify-ing variable speed pumps, the userrequired only three pumps: one foreach duty level and a single sparewhich was valved to allow opera-tion under either condition. Furthersavings were achieved for the mainpumps since identical pumps wereused (desired conditions were metby varying the speed). Parts wereinterchangeable and significantlyless energy was required when run-ning Pump A at the normal (i.e.,low-head) condition.

In addition to covering a widerange of conditions, variable speedpumping can also eliminate theneed for multiple stages. With

increased speed, centrifugalpumps produce increasedhead and flow.

As mentioned above,variable speed pumping canalso eliminate the need fora throttling valve. Also,bypass valves may nolonger be necessary sinceminimal flow requirementsfor stable operationdecrease with speed.Elimination of valves canreduce capital expense,maintenance costs, risk ofleakage and pressure losses(pressure drop across thevalve often accounts for 10percent of total pressurerise required).

One user saved $20,000 byconverting to variable speed

pumping in an application involvinginjection of water into the combustionchamber of gas turbine engines. Sincethe system curve had relatively littlestatic lift, the pump could be slowed toproduce only the desired flow and headand still maintain good efficiency. Achange from a fixed speed pump withthrottling valve and bypass valve tovariable speed eliminated the twovalves, reduced the power requirementof the system from 100 hp to 75 hpand made the assembled skid ofequipment smaller.

Dramatic power savings are avail-able because of reduced head and

flow points due to changing speedrather than by dis-charge throttling(Figure 4). For instance, by achieving60 percent of design flow and headthrough variable speed, users cansave 50 to 80 percent on energy costscompared to fixed speed pumpingwith a throttling valve.

Another advantage variable speedpumping offers is reduced heat to thepumped fluid. At constant speed,efficiency falls with reduced flowrate. The result of hydraulic ineffi-ciencies is heat rise in the fluid. Butvariable speed pumps remain effi-cient at low flows (i.e., low speeds).Furthermore, horsepower levels arelower at low speeds, which meansthat heat input to the fluid is keptminimal. Variable speed pumpingcan thus be advantageous for lighthydrocarbon and other volatile fluidapplications.

SELECTING THE RIGHT SIZE PUMP

Like any pumping application,variable speed pumping requiresproper sizing of pumps. But unlikeconstant speed pumps, variablespeed pumps are not selected for asingle design point. To select the cor-rect size pump, you should constructthe desired head versus flow rangefor all anticipated specific gravities.Then be sure to specify a pump thatcan cover that range (Figure 5 showsa pump that cannot reach point B).

Variable speed centrifugal pump operation

FIGURE 2

Effects of changing speed of acentrifugal pump

FIGURE 3 FIGURE 4

Hydraulic HP savings for a centrifugalpump

FLOW PROPORTIONAL TO SPEED - NO STATIC LIFT

Page 56: Centrifugal Pumps Handbook

The Pump Handbook Series 57

VFDs. High efficiency is not arequirement, but the extra copperand other features are advanta-geous for VFD use.

Increased heat can lead to envi-ronmental hazards. Motors pro-posed for use in hazardous (e.g.,explosive) environments must bedesigned differently or derated.The skin temperature of a standardmotor operating on a VFD couldexceed an area gas autoignition temperature at nameplate horse-power. Motors nameplated for usein Class I, Division I, Groups C andD environments, for example, areavailable for VFD use but mustgenerally be purchased with a“matched” VFD from a single sup-plier.

SELECTING A VFDImportant factors for selecting

VFDs include power supply voltageand frequency, amperage require-ments, torque requirements andmotor and load characteristics.

VFDs must be selected to matchthe power supply and frequency.Many VFDs are switch selectablefor a number of voltage/frequencycombinations.

You can determine the amper-age requirement of a motor usingthe equation:

You may need to specify a “fictitious”100% speed point to ensure thepump has adequate range (Figure 6).

You must also ensure thatNPSHA and motor horsepower areadequate for all combinations offlow and speed. NPSHR and efficien-cy vary approximately as the square of the speed (Figure 7). Since NPSHRincreases with speed, in-ducers maybe required to reduce NPSHR to avail-able levels. Bearing loads and otherpump characteristics must also becarefully examined.

MOTOR-VARIABLE FREQUENCY DRIVEBEHAVIOR

One of the most common meth-ods of changing motor speed is theAC Variable Frequency Drive (VFD).VFDs are designed to take advantageof the fact that speed, torque andhorsepower of an AC motor are allrelated to the frequency and voltageof the electric power supply:

Torque Capability = F(volts/hz)

HP Capability = f(Torque x Speed)

VFDs convert incoming AC elec-trical power to DC then invert theDC power into variable frequencyand voltage AC power. A numberof technologies are available toswitch the DC power through semi-conductors to achieve the desiredvoltage or current pulses. The tech-nologies differ in their ability tocreate optimal waveforms. Becausethe motor’s torque and torque rip-ple are determined by the current,the VFD affect motor and pumpoperation. Thus, by knowing thecharacteristics of the VFD output,you can select a VFD suitable foryour pump.

Most VFDs produce a constantvolt/hz ratio, thus constant motortorque capability up to name-platefrequency (typically 60 hz or 3550rpm for a two-pole motor — seeFigure 8). Horsepower capabilitytherefore rises from zero at zerospeed to full horsepower at name-plate speed. Above nameplate speed,

2 x hz x 60 # of Poles

Nominal speed

the VFD cannot provideincreasing voltage, so torquefalls due to the falling volts/hzratio. Horsepower capability,however, remains constantsince speed is increasing.Electrically, induction motorscan be run at approximately90 hz in this configuration.But mechanical constraintsmay limit the safe runningspeed to well below 90 hz.

VFDs can be used to pro-vide extra motor horsepowerabove 60 hz. Recall that motortorque capability is propor-tional to the volts/hz ratio. If amotor is designed for a givenvolts/hz ratio, and that ratiocan be maintained at a higherspeed, torque capability will be constant.This technique can frequently beused with standard motors which arecommonly wound for either 230 V or460 V at 60 hz. By connecting for 230V at 60 hz and operating to 460 V at120 hz, both motor and horsepowercapability and speed are doubled. Besure to check with the motor manu-facturer before using this technique.The motor may not have the thermalcapacity or mechanical integrity torun at speeds considerably above 60hz. Also, the motor may not be prop-erly matched electrically to the VFD.

SELECTING THE MOTORVFDs are most frequently used with

the familiar NEMA B squirrel cage ACinduction motors. Some specialconsiderations for selecting motorsfor use with VFDs include cooling,efficiency and operation in haz-ardous (e.g., explosive) environ-ments.

Motors operated on VFDs oper-ate at higher temperatures due tothe irregular shape of the electricalwaveforms produced by the VFD.To ensure that the motor will notoverheat, the motors are typicallyderated at full load from 3 to 10percent, depending on the type ofVFD used.

This additional heat makesmotors operated on VFDs less effi-cient than when operated acrossthe line. Thus, many users specifyhigh efficiency motors for use with

100% SPEED-PUMP 1

LIMITS OFCAPABILITYPUMP1

CAN'TDO

"B"

"A"BEPDESIRED

RANGE OF HEAD & FLOW

120

100

80

60

40

20

00 20 40 60 80 100 120 140

FLOW%PUMP 1 SIZED FOR "A"-UNABLE TO DO "B"

HE

AD

%

FIGURE 5

Improper sizing to meet required duty points

Page 57: Centrifugal Pumps Handbook

58 The Pump Handbook Series

9). Fortunately, many pumps areof a stiff staff design and will oper-ate below their first lateral criticalspeed. A vendor may be able tochange the mechanical design toraise or lower the critical speed toprovide full range speed adjust-ment.

Be aware of torsional criticalspeed. Torsional critical speeds areresonant frequencies at whichmotor and driven equipment shaftscan begin to oscillate with angulardisplacement as a result of torsion-al excitation. VFDs can cause tor-sional excitation problems knownas torque ripple. For example,rather than delivering a continuous295 ft-lb of torque, a VFD-driven,200 HP motor may deliver torquecycling between 250 and 340 ft-lbat some 21,000 cycles per minute.This oscillation could be damaging.Clearly, careful analysis and selec-tion of the VFD, motor, couplingand pump train are needed toavoid torsional problems.

ENVIRONMENTAL CONSIDERATIONSTo avoid potential problems in

your application of VFDs, youmust take a few precautionsregarding their environment.

Locate VFDs indoors. Units canbe placed outdoors with the prop-er enclosure, but the cost of theenclosure can run into thousandsof dollars. Fortunately, the VFDcan be up to several hundred feet

AMPS =HP x 746

Volts x 1.732 x Motor Efficiency x Motor Power Factor

Nominal horsepower ratings areusually given by the VFD vendorsbut in some instances a VFD willonly produce the stated nominalhorsepower if a high efficiencymotor is used. Unlike motors,VFDs generally have no continuousservice factor. Momentary over-loads, however, are permitted.VFDs generally exceed 97 percentefficiency at full load.

VFDs are designated constanttorque or variable torque, depend-ing on their current overloadcapacity. Variable torque VFDs canproduce 110 percent of rated cur-rent for one minute. Constanttorque VFDs can produce 150 per-cent of full load current for oneminute and even more for shorterperiods. Variable torque VFDs aregenerally used for centrifugalpumps.

VFDs must be matched to theload and motor characteristics.Certain VFDs, known as CurrentSource Inverters or CSIs, may requireaddition or deletion of capacitorbanks to match the load and motor

characteristics. Themore commonly usedPulse Width Modul-ation and Six StepVFDs do not requirethis matching. Theyare suitable for awide variety ofmotors. Most VFDsoperate on 480 Vinput and produce amaximum of 480 Voutput. If a highervoltage motor isdesired, you caninstall a step-up trans-former between theVFD and the motoror use a higher volt-age VFD.

APPLICATIONCONSIDERATIONS

Be sure the motorwill be capable of delivering enoughtorque to the pump. Motor torquecapability (including breakaway orstart-up torque) must exceed pumptorque required at every speed.Generally, if the motor and VFD areproperly sized for 100 percent speed,they will be adequate at lowerspeeds. However, in certaininstances, such as applications withhigh suction pressure, motor andVFD sizing may be governed bystart-up conditions. VFDs on positivedisplacement pumps must routinelybe oversized to provide sufficientstart-up torque.

Avoid lateralcritical speeds. Asan example, APISpecification 610states that depend-ing on the unbal-anced responseamplification fac-tor, a pump maynot be operatedbetween 85 percentand 105 percent ofits critical speed.Adherence to theserules can block outa large portion ofthe allowable per-formance envelopeof the variablespeed pump (Figure

EFF

ICIE

NCY

@

80%

SPEED

EFF

ICIE

NC

Y@

100%

SPEED

EFF

ICIE

NC

Y@

120%

SPEED

NPSH @ 120% SPEED

NPSH @ 100% SPEEED

NPSH @ 80% SPEED

200

150

100

50

0

60

50

40

0 25 50 75 100 125 150

NP

SH

R

EF

FIC

IEN

CY

FLOW - % OF 100% SPEED FLOW

Centrifugal pump NPSHR and efficiency vs. speed

100% SPEED-PUMP 2

78% SPEED-PUMP 2

LIMITS OFCAPABILITYPUMP2

"B"

"A"

"C"DESIREDRANGE OF HEAD & FLOW

120

100

80

60

40

20

00 20 40 60 80 100 120 140

FLOW%PUMP 2 SIZED FOR "C"-UNABLE TO DO "A" & "B"

HE

AD

%

68% SPEED PUMP 2

Proper sizing to meet required duty points

FIGURE 6

FIGURE 7

Page 58: Centrifugal Pumps Handbook

The Pump Handbook Series 59

from the motor. So it can beindoors even if the motor is out-doors.

Derate for high temperatures andhigh elevations. If operated above104o F, VFDs must be derated. Theymust also be derated if used at eleva-tions above 3300 ft.

Be cautious of power supply.VFDs are sensitive to stiffness andirregularities in the electrical sup-ply. You may need to install a linereactor or isolation transformerbetween the VFD and supply mainif the feed transformer is very stiff(high KVA). Input line reactors orisolation transformers may also benecessary to prevent the VFD fromfeeding electrical noise back into thesupply main. Such noise can distort

instrument signals ifthey are fed from thesame supply trans-former as the VFD.

IS IT WORTH IT?Despite the list

of precautions, vari-able speed pumpingcan save youmoney. As shown,you can eliminatethe need for throt-tling valves. Youmay be able to useone variable speedpump in place of twofixed speed pumps.

VFDs also elimi-nate the need for a

motor starter. Variable speedpumping often reduces powerrequirements. And some electricalutilities provide rebates for compa-nies that use energy saving devicessuch as VFDs. Rebates can be up toone-third the purchase price of thedevice. Other cost savings comethrough better process control dueto lower heat inputs and fluidshear.

These savings frequently payback the costs of utilizing variablespeed pumping (such as the cost ofthe VFD, possibly extra costs forhigh-efficiency motors and possiblyoversized pumps). Payback periodsof as little as one year are typicalwhen using variable speed pump-

ing.

THE FUTUREVariable speed pum-

ping will become morepopular as the technol-ogy establishes itstrack record. And asmore system and plantengineers design forvariable speed opera-tion early in the devel-opment cycle, benefitsbeyond energy conser-vation will becomeapparent.

Advances in VFDtechnology will alsoincrease user accep-tance. New featuressuch as greater adjust-

ment in operating parameters willmake VFDs easier to use and inte-grate into a system.

Improved reliability and faulttolerance will make VFDs easier toapply. You can expect manufactur-ers to add adjustment capabilitiesof output voltage and currentwaveforms to optimize motor effi-ciency and smoothness. Improve-ments in power semiconductorswill provide higher efficiency andsmoother output.

Sizes of VFDs will diminish ascomponents on circuit boards areintegrated into chips. Reduced sizeand improved efficiency will allowpackaging to be more compact andenvironmentally rugged, whichwill allow placement even in haz-ardous environments.

Prices will come down, possi-bly by up to 25 percent over thenext five years.

Even today, you can achievegreater flexibility, energy savings,equipment savings and extra headand flow through variable speedpumping, provided you take extracare in assembling an appropriatecombination of pump, motor andVFD. With improvements in tech-nology, more and more users willbegin to take advantage of variablespeed pumping. ■

Stephen P. Murphy is SeniorBusiness Development Specialistfor Sundstrand Fluid Handling inArvada, CO.

Performance of conventional variable speed motor

POSSIBLETHERMALDERATING

T

T

HP

HP

0 30 60 90 1200 30 60 90 120

100%

50%

460

230

%T,%HPV

HZ HZT = f (V/HZ)

HP = f (TXSPEED)

FIGURE 8

120

100

80

60

40

20

00 20 40 60 80 100 120

FLOW

100% DESIGN SPEED

50% DESIGN SPEED

}PREDICTEDCRITICAL SPEED

CRITICALSPEEDAVOIDANCEBAND

HE

AD

%

5%

15%

Avoiding lateral critical speeds

FIGURE 9

Page 59: Centrifugal Pumps Handbook

60 The Pump Handbook Series

ith greater global competi-tion and increased environ-mental regulations, mod-ern industrial applications

over the years have evolved intosophisticated operations, demandingmore control over their liquid han-dling processes. This is particularlyevident on the ”dirty“ liquid side of aplant’s manufacturing process, in thedrainage, filtration/pollution con-trol/wastewater areas. Self-primingcentrifugal pumps are important inmeeting this demanding challenge.

Single stage end suction centrifu-gal pumps may be divided by theirdesigns into conventional or standardcentrifugals and self-priming centrifu-gals. Centrifugal pumps incorporate asimple design with minimum movingparts - impeller, shaft and bearings.They are reliable, durable and rela-

tively easy to maintain. To betterunderstand the working principle ofa self-priming centrifugal pump, let’sfirst examine the centrifugal forceprinciple and a standard or conven-tional centrifugal pump.

All centrifugal pumps incorpo-rate the centrifugal force principle,which may be illustrated by a carrunning on a wet road (Figure 1).The tires pick up water and throw itby centrifugal force against the fend-er. Centrifugal pumps incorporatethe same principle, but the tire isreplaced by an impeller with vanesand the fender is replaced by thecasing (Figure 1b). The liquid entersthe center or eye of the impeller. Asthe liquid reaches the impeller vane,its velocity is greatly increased.Centrifugal force, created by theimpeller blades or vanes, directs the

liquid towards the outside diameterof the impeller. Once the liquidreaches the tip of the impeller vaneit leaves the impeller at its greatestvelocity. As the liquid leaves theimpeller, its direction is controlledby the pump casing (the most com-mon casing shapes are spiral orvolute and circular).

The spiral or volute casing sur-rounds the impeller, beginning at thepoint where the liquid leaves theimpeller. The liquid enters the casingand follows the rotation of theimpeller to the discharge. Within thecasing there is a section called thethroat or cutwater.

The cutwater, also called thetongue, is a cast section of the volutecasing, near the discharge that ispositioned close to the maximumimpeller diameter. As the liquid

BY TERRY W. BECHTLER

W

Self-Priming Centrifugal Pumps

The ability to self-prime can be a cost effective solution for many applications.

FIGURE 1

A B

CENTRIFUGAL PUMPS

HANDBOOK

LOWPRESSURE

Page 60: Centrifugal Pumps Handbook

The Pump Handbook Series 61

reaches the cutwater it is divertedinto the pump’s discharge opening(Figure 2).

SELF-PRIMING

Self-priming centrifugal pumpsincorporate all the above standardcentrifugal pump design features andadd the following internal modifica-tions:

• A casing design that surroundsthe volute and impeller andenables the pump to retain liquidin a built-in reservoir, or primingchamber. This reservoir is filledduring the initial prime of thepump, and when the pump com-pletes a pumping cycle and shutsdown, the reservoir retains liquidfor the next priming cycle.

• An internal recirculation channelor port. This channel connectsthe pump’s discharge cavity backto the suction reservoir internal-ly, allowing the continuous recir-culation of liquid from dischargeback to suction during the prim-ing, usually to the peripheral por-tion of the impeller (Figure 2B).

These two internal design fea-tures, the priming chamber and inter-nal recirculation channel, are whatdistinguishes a self-priming centrifu-gal pump from a standard centrifugalpump. Self-priming can also beaccomplished by a diffuser designcentrifugal pump that is used primari-ly for clear liquids.

HOW IT WORKS

Self-priming centrifugal pumpscan be placed above the liquid levelof the source (Figure 3). Only thesuction pipe enters the liquid beingpumped. The pump is initiallyprimed by adding liquid to the pumpcasing through a priming port, nor-mally located near the discharge. Theliquid fills the discharge reservoir,traveling into the eye of the impellerthrough the pump’s recirculationchannel. The suction line, itself, isnot filled. A check valve is usuallylocated just inside the suction reser-voir. All connections must be air-tight. During initial start-up, theimpeller rotation causes the liquid

in the pump reservoir to be direct-ed to the discharge cavity via cen-trifugal force. Simultaneously, alower pressure is formed in the suc-tion reservoir. This draws the liquidfrom the discharge cavity back intothe suction reservoir through thepump’s internal recirculation chan-nel. This is a continuing action dur-ing the priming cycle. While this isoccurring, the air in the suction lineis drawn by the lower pressure intothe eye of the impeller with thepriming liquid and travels throughthe volute into the discharge cavity.At this point velocities decrease,allowing the air and liquid mixture toseparate. The air flows up and isejected, and the priming liquid recir-culates back into the impeller.

This process continues to drawall the air from the submerged suc-tion line. In applications where theliquid level is at atmospheric pres-sure, that pressure on the liquid sur-face, coupled with the lower pressurein the suction pipe due to the evacua-tion of air, serves to push the liquidin the sump into the pump. When allair is evacuated liquid pumping auto-matically begins.

Note that the diffuser design self-prime principle incorporates animpeller rotating in a stationarymulti-vane diffuser (Figure 4). Duringpriming, the diffuser separates the airfrom the pumped liquid until primingis completed.

This priming action might seemsomewhat complicated or mysteri-

ous, but it is actually a very easy taskfor a correctly installed self-primingcentrifugal pump, and it happensautomatically in a relatively shorttime (20 - 30 seconds for a normal 15foot suction lift).

It’s this feature that differentiatesself-priming centrifugal pumps fromstandard centrifugal models. On asuction lift condition, a standard cen-trifugal pump, with only air in thecasing and having no ability to sepa-rate air and liquid to create a vacu-um, would have an impeller thatsimply spins, acting as a fan, becauseit has no way to lower the suctionline pressure. By placing a foot valveon the end of a suction line and fillingthe pump and suction line with liq-uid, a standard centrifugal pump canbe made to operate and pump in aconventional mode.

If the foot valve leaks and airenters the suction, such as under ashutdown condition, a standard cen-trifugal pump stands the risk of losingits prime and becoming air bound.Under suction lift conditions, self-priming centrifugal pumps are idealfor unattended use.

Standard centrifugal pumps aresometimes fitted with priming sys-tems to fill the pump and suction linewith liquid prior to starting. In suchcases, a control device tells the pumpwhen all air is evacuated and the unitis liquid filled to start.

STYLES

Self-priming centrifugal pumpsare usually classified into two groups:basic self-priming pumps and trash-handling self-priming pumps.

Basic self-priming pumps usuallycome with different impeller configu-rations, including fully enclosed andsemi-open. Like all centrifugalpumps, the pressure developed isdependent on the impeller diameterand rpm.• Fully enclosed impellers allow

self-priming pumps to developmedium to medium-high dis-charge pressures, up to about 110psi or 254 ft total dynamic head(TDH). Normal pump sizes rangefrom 1 in. through 6 in. suctionand discharge. Pumps with afully enclosed impeller have avery limited solids handling capa-

FIGURE 2

Page 61: Centrifugal Pumps Handbook

62 The Pump Handbook Series

bility, with sizes from 1 1/32 in.through 5/8 in. in diameter,depending on the size of thepump. This configuration isexcellent for handling clear liq-uids, including processed hydro-carbons, along with generalwash-down pressure applications.

• Semi-open multi-vane impellersare usually designed for slightlylower head conditions than fullyenclosed impellers, but theyhave greater solids handlingcapabilities. Pump sizes usuallyrange from 3/4 in. through 12 in.suction and discharge, withcapacities to more than 5,500gpm. Spherical solid sizes rangefrom 3/4 in. through 3 in. indiameter, depending on the sizeof the pump. Basic self-primingpumps with semi-open impellers

are sometime referred to as gen-eral-purpose self-priming pumps.They are excellent for handlingdirty, contaminated liquids.Applications include extensiveuse in industrial filtration opera-tions and a wide range of engine-driven models that serve theconstruction market.

Trash handling self-primingpumps generally use a trash-type,semi-open, two-vane impeller thatallows the pump to pass larger spher-ical solids.

• Trash handling self-primingpumps generate medium dis-charge pressures in the area of62 psi or 145 ft TDH on electricmotor drives and discharge pres-sures upwards of 75 psi or 173 ftTDH on engine-driven configu-

rations, with capacities upwardsof 3,400 gpm. Normal pumpsizes range from 1-1/2 in.through 10 in. suction and dis-charge. The impeller designallows for excellent solids han-dling capability, ranging from lin. to 3 in. spherical solids diam-eter, depending on the pumpsize.Trash handling self-priming

pumps are often referred to as theworkhorse of centrifugal pumps dueto their rugged design and largesolids handling capabilities. Thesepumps can be found on some of themost severe pumping applicationswithin plants or on constructionsites.

A desirable design feature of atrash handling self-priming pump is aremovable cover plate, located

A cut-away view of a self-priming centrifugal pump designed to handle solids-laden liquids andslurries

FIGURE 2B

DischargeOutlet

Flap Valve

SuctionInlet

ReplaceableWearplate

Removable Coverplate

PressureRelief Valve

BalancedImpeller

Volute

Ball Bearings

CartridgeMechanical

Seal

Page 62: Centrifugal Pumps Handbook

The Pump Handbook Series 63

directly in front of the impeller onthe suction side of the pump. Trashhandling self-priming pumps may beapplied in waste sump applicationswhere they are exposed to varioussize solids. Any pump may clog try-ing to pump larger solids than it wasdesigned to pass. The removablecover plate allows quick access to thesuction side of the pump, expeditingthe removal of blockage. Somedesigns allow removal of the coverplate without disturbing the suctionand/or discharge line.

SELECTIONAs discussed, self-priming cen-

trifugal pumps have a broad designrange that allows them to serve awide variety of applications. Manymetallurgical choices and shaft sealconfigurations are available to bestserve particular services.

Mechanical shaft seals can besingle, double, or tandem. They areavailable as double grease lubricatedfor general purpose applications, oillubricated with silicon carbide facesfor industrial applications with abra-sives, carbon against Ni-Resist facesfor clean water or refined hydrocar-bon applications, or Teflon fittedwith carbon/ ceramic faces for chem-

ical applications. Alloysavailable for pumpconstruction also offerthe same diversity.Cast iron and ductileiron are used for gener-al purpose and refinedhydrocarbons, hard-ened austempered duc-tile iron (ADI) isemployed for abrasiveapplications, CD4MCuSS serves in corrosiveand abrasive applica-tions, and 316 SS, Alloy20 SS, Hastalloy B, andHi-Resin Epoxy Plasticare used for other spe-cial chemical applica-tions.

APPLICATIONGUIDELINES

The principalapplication area forself-priming pumps iswhere their ability toself-prime is a costeffective solution; and

when it is more convenient and desir-able to locate a pump ”high and dry“above the liquid. Some general guide-lines are in order:

The liquid being pumped shouldbe of low viscosity ( 550SSU or less).Horsepower and efficiency correc-tions are needed for liquid viscosityabove 550 SSU. If subjected to liquidfreezing temperatures, the pump

must be protected against freezing toavoid damage.

The vapor pressure of the liquidand the presence of high levels ofentrained air are serious considera-tions in suction lift application.

The NPSHA (net positive suctionhead available) must exceed the man-ufacturer’s published NPSHR (netpositive suction head required) by amargin that accounts for the liquidproperties.

Repriming time increases withsuction lift. Suction lifts with water asthe liquid at normal ambient temper-ature should be limited to 15 to 18 ft.best efficiency range, although maxi-mum practical lifts are obtainable to25 feet. For other liquids or liquidmixtures, the vapor pressure of theliquid or the most volatile compo-nents of a mixture must be consid-ered. Reducing the speed of operation(rpm) significantly reduces theNPSHR. Suction line piping shouldbe sized to velocities in the 5 to 7 ft.range at design flow. For self-primingpumps it is recommended that thesuction piping should be the samesize as the pump’s suction inlet.

The self-priming centrifugalpump offers a unique solution tomany pumping applications. ■

Terry W. Bechtler has beenManager of Inside Sales for TheGorman-Rupp Co. in Mansfield, OH forfour years.

FIGURE 3

FIGURE 4

Volute Priming Diffuser Priming

Page 63: Centrifugal Pumps Handbook

64 The Pump Handbook Series

Centrifugal Pump TestingLaboratory and on-site testing ensure pumps are up to their tasks.

BY LEO RICHARD

s industry becomes increas-ingly competitive, pumps arebeing sized to precisely meettheir duty requirements with-

out oversizing. This allows users tomaximize efficiency and minimizefirst capital costs. There is also asmall but growing trend to questionthe economics of in-line spares andlarge spare parts inventories. Thesedevelopments make it more criticalthan ever that rotating equipmentprecisely meets all hydraulic, materi-al, and safety requirements. This isassured by thorough testing of perti-nent parameters by manufacturersprior to shipment and by customersat their job sites.

The level of justifiable testingwill depend on the nature of the ser-vice and significance of the parame-ter to be measured. For instance, awater transfer application can beserved with a stock pump that hasundergone the manufacturer’s stan-dard quality and performance checks.However, a corrosive, high pressure,or environmentally hazardous appli-cation may justify additional testingfor material conformity and quality ofconstruction.

In addition to the extent of test-ing, several other factors must also beconsidered. The first is location. Ashop or laboratory test is typicallyconducted at the manufacturer’s facil-ity. The test lab provides a tightlycontrolled environment and therebygenerates the most accurate data. Incontrast, field tests sacrifice someaccuracy, but they provide usefuldata under the actual conditions ofservice.

A reasonable split between thetwo approaches should be employed,depending on the nature of the evalu-ation and the user’s ability to conducton-site testing. Also, the user andmanufacturer must agree to a set ofguidelines such as those published bythe Hydraulic Institute (HI). Amongother things, HI standards generallydefine the methods and acceptabletolerances to be used. However,regardless of the standard employed,good laboratory practice requires thatall instrumentation be calibratedprior to the test. For maximum accu-racy the instruments should be locat-ed after straight runs of pipe wheresteady flow conditions exist. In addi-tion, the local barometric pressuremust be considered, especially inapplications requiring suction lift.The data obtained should be recordedin a test log, and each round of evalu-ations must be identified in this docu-ment by the manufacturer’s anduser’s serial/ equipment numbers.

Also, the question of user repre-sentation during testing should beclearly defined. This includes issuessuch as site location, the amount ofadvance notice prior to testing, andcost. Usually, the added cost and logis-tics problems make such witness test-ing inadvisable—unless the user hasvery limited experience with the man-ufacturer. This, as well as any otherrequirements, must be written into thespecification prior to purchase.

A brief description of the mostcommon performance and qualityevaluations is given below. For sim-plicity, these tests have been charac-terized in terms of certifying

AThe test lab

provides

a tightly

controlled

environment

and thereby

generates

the most

accurate data.

CENTRIFUGAL PUMPS

HANDBOOK

Page 64: Centrifugal Pumps Handbook

The Pump Handbook Series 65

conformance in hydraulic capability,materials, or physical integrity.

HYDRAULIC CAPABILITYThe determination of hydraulic

performance is the most basic andcommon category of testing. Thistypically involves performance, netpositive suction head (NPSH), andpower evaluation.

PERFORMANCE TESTINGThe performance test of a spe-

cially ordered or job pump typicallyinvolves the generation of its head-versus-capacity curve at the ratedimpeller trim. Such pumps are shoptested on water at the manufactur-er’s site. If the HI standards are fol-lowed, the acceptance level must bedefined. Level A requires that seventest points of head, flow, and effi-ciency be evaluated. Level B testingrequires that five test points bechecked. Each level of acceptancerefers only to the head and capacityas specified by the customer for theservice, also known as the rated orguarantee point. The defined toler-ances for these parameters will varydepending on the size of the pumpsand the level of testing required.

NPSHThe NPSH test is basically a

measure of the suction head require-ment necessary to prevent cavitation.The procedure typically involves

holding the flow constant and reduc-ing the suction head until a definedlevel of cavitation occurs. The dataare used to generate a curve of thecavitation coefficient, Sigma, for thepump at the specified capacity. Sigmais defined as the net positive suctionhead available divided by the totalpump head per stage. According to HIstandards, the NPSH requirement ofthe pump is defined as the point atwhich a 3% head drop occurs on theSigma curve. However, this criteria issomewhat controversial. The majorissue is that incipient cavitation is wellunder way prior to the occurrence ofthe 3% head drop. In fact, some com-panies are considering an internalspecification defining the NPSHrequirement as only a 1% head dropon the Sigma curve.

POWER/EFFICIENCY TESTINGPower and efficiency testing is

becoming increasingly critical ascompanies are closely evaluatingpower consumption during the pumpselection process. Another relevantissue is the growing trend of retro-fitting sealed applications with seal-less designs. As is well known, due tomagnetic coupling and viscous losses,sealless pumps inherently have slight-ly greater power requirements thantheir sealed counterparts. Therefore,a confirmation of the publishedpower requirements may be in order,especially for installations where

existing motors and starters are to bereused. Such tests are typically con-ducted on water using certifiedmotors. Data are collected at severalpoints, depending on the level speci-fied as part of the performance test.This information can be used to gen-erate both wire to water andhydraulic efficiencies.

MATERIAL CONFORMANCEThe usual considerations for

material conformance testing are cor-rosion and erosion resistance. Again,the added cost for these proceduresmust be justified with regard to theparticular application, as well as theconsequences of process downtimeand personnel or environmentalexposure.

CERTIFICATE OF MATERIALCONFORMITY

The most basic type of documen-tation is in the manufacturer’s certifi-cate of material conformity. This is aguarantee that the pump is made ofthe materials called out in the specifi-cation. This certificate is based solelyon the standard quality tests per-formed by the manufacturer.

CHEMICAL ANALYSISThis involves confirmation of the

material of construction by chemicallytesting small samples from the pump.These tests range from sophisticatedchemical analysis to a basic screeningutilizing chemical test kits.

NUCLEAR ANALYSISThis confirms materials used by

means of a nuclear analyzer. This is anondestructive test involving directmeasurement on the surface to beanalyzed. The composition of thematerial is determined by the equip-ment and matched with its internaldatabase to generate an identification.Due to the high cost of this equip-ment, many sites utilize sub-contrac-tors for this work.

HARDNESS TESTINGHardness testing may be

required by the user, especially forpumps in highly erosive services.Though the type of hardness test canvary, the Brinell hardness test is fairlycommon.

A technician attaches a mag drive pump to a test tank.

Page 65: Centrifugal Pumps Handbook

66 The Pump Handbook Series

MILL CERTIFICATIONAn extensive form of evaluation

involves mill certification. Basically,the mill certs follow the pump alongeach step of the manufacturingprocess. This includes data from theinitial pour at the foundry to the finalchecks of the finished components.The downside of mill certification isthat it tends to be costly. Also,because additional data are requiredfrom the initial pour, stock pumpsmay not be used. Some parameterstypically measured in mill certsinclude:

•Mechanical Test Certification,which includes tensile strength,proof stress, and elongation.

•Analysis certificates detailing thechemical composition.

•Intercrystallation corrosion andultrasonic tests.

PHYSICAL INTEGRITY TESTINGAs the name implies, this catego-

ry of testing basically involves a con-firmation of the pump’s ability tomaintain the liquid boundary underthe conditions of service. The chiefareas of concern prompting such test-ing are the integrity of welds and pos-sible porosity of castings.

DYE PENETRANT TESTINGDye penetrate testing

involves the use of anextremely low surface ten-sion liquid to detect possi-ble leak paths in cast andwelded surfaces. If the dyepenetrates the surface, thepiece is either rejected orweld repaired. If the com-ponent is repaired, the useris notified and the partretested to confirm theintegrity of the weld.

RADIOGRAPHYRadiographic testing is

primarily used to confirmthe integrity of welds inpressure-containing compo-nents. Procedures dependon the configuration and dimensionsof the component, as well as thenature of the equipment being used.The test itself is somewhat costly andmay impact delivery. For these rea-

sons its use is most often limited tocritical applications in the powerindustry.

GAS LEAK DETECTIONThis involves pressurizing the

pump with an inert gas such as arc-ton to detect any leak paths from thepump. Leaks are typically detectedby means of a sniffer or mass spec-trometer. This test is extremely sensi-tive and able to detect the slightestporosity in castings.

HYDROSTATIC TESTINGHydrostatic pressure testing is a

standard quality check. The proce-dure usually involves filling pressure-containing components with waterand pressurizing to 1.5 times therated working pressure. This pressureis held for a specified time, and thepiece is inspected for leaks.

TESTING SEALLESS PUMPSThe testing procedures utilized to

evaluate standard sealed centrifugalsare commonly used for sealless con-figurations as well. However, due tothe unique design of sealless pumps,some additional procedures may beconsidered. A complete discussion ofthis topic can be found in theHydraulic Institute Standard for seal-less centrifugal pumps. (HI 5.1–5.6,1st edition, 1992)

MAGNETIC STRENGTHThe strength of the

permanent magnets in amagnetically driven seal-less pump can be evaluat-ed with a Gaussmeter.This instrument directlymeasures the strength ofthe magnetic field inGauss or Milligauss.Gauss testing is usually anoverkill for new pumpsfor the following reasons:

• The relative unifor-mity of productionmagnets.

• The high safety fac-tor incorporated into a

magnetic coupling’s power trans-mission capability. (A safety factorof 2.0 under full load conditions istypical.)

• The fact that the coupling isalready inherently tested duringgeneration of the hydraulic perfor-mance curve.

BREAKING TORQUEA “low tech” but effective way of

site testing synchronous magneticcouplings is by measuring the break-ing torque. Breaking torque is simplythe force required to break or decou-ple the two opposing halves of themagnetic coupling. This is accom-plished by anchoring the inner rotat-ing assembly and applying torque tothe outer magnet ring (OMR). Force isapplied and measured by a torquewrench fitted to the drive shaft of thepump (the drive shaft is mechanicallycoupled to the OMR). The data gener-ated is then compared to the manu-facturer’s standards.

As in Gauss testing, this proce-dure is usually unnecessary for a newpump. However, it is a useful fieldtool for confirming the strength of themagnetic coupling. This is especiallyimportant during a rebuild after a dryrun failure. During dry runs, the mag-nets are exposed to extreme tempera-tures that may reduce their strength.By utilizing the breaking torque pro-cedure, maintenance personnel canpretest the magnetic coupling prior toreinstallation.

SECONDARY CONTAINMENT TESTING

CANNED MOTOR DRIVESThe stator housing in canned

motor pumps is often used as a sec-ondary containment vessel. Testingtypically involves gas leak detectionon the finished stators. For designsutilizing potting of the wire leads,confirmation of the integrity of thesecondary containment chamber asthe equipment ages may be in order.This is especially relevant in serviceswith high temperature cycling, whichmay damage the potting compound.

MAG DRIVE DESIGNSIn some mag drive designs the

coupling housing and an inboardmagnetic seal are utilized for sec-ondary containment. Testing usuallyinvolves a hydrostatic or gas leakdetection of this assembly.

Thedeterminationof hydraulicperformance is the mostbasic andcommoncategory

of testing.

Page 66: Centrifugal Pumps Handbook

The Pump Handbook Series 67

SITE TESTINGOne of the most important and

often overlooked opportunities forevaluating and documenting pumpperformance is the initial commis-sioning. Information gathered at thistime is critical in verifying initial per-formance and providing a bench-mark for future diagnostic andtroubleshooting efforts.

It is suggested that, as a mini-mum, the following areas be evalu-ated:

• The total differential head generat-ed by the pump. It is strongly rec-ommended that both suction anddischarge gauges be installed tofacilitate measurement of thisparameter. Once determined, itshould be noted whether the actu-al operating point differs from theduty listed in the specification. Ifso, the user must first confirmproper operation of the pump andprocess. If these check out, anevaluation of potential problemsassociated with the new dutypoint must be evaluated. Thisincludes a possible increase in theNPSH requirement and powerconsumption. Also note that con-tinuous operation at extremelyhigh or low flows will significantlyincrease dynamic loading on the

impeller. Such loading can dramat-ically decrease the mean timebetween failures for the equip-ment.

• Evaluation of the operating pointshould be conducted for all condi-tions the pump will experience.For example, many pumps intransfer applications deliver liquidto various locations and are peri-odically operated in a recirculationmode. Each of these duty pointsmust be determined and possibleproblems identified. If necessary,modifications in the pump and/orprocess should be made. Commoncorrective actions include resizingorifices, changing valve settings,and adjusting the impeller trim.

• The amp draw of the motorshould be measured. This is thencompared with the manufacturer’sstated requirements to evaluateproper operation. Gross differ-ences between these figures mayindicate various conditions such ascavitation, operating to run out orshut in, or mechanical problems.

• The vibration level should be mea-sured. This will confirm properoperation and serve as a bench-mark for future testing.

• After the pump has achievedsteady state, the bearing frame,process, and ambient tempera-tures should be monitored andrecorded. This data will be usedas an initial check as well as forfuture reference.

• Proper operation of all protectiveinstrumentation should also beverified and any outputs record-ed. For instance, many seallesspumps utilize a thermocoupletemperature monitoring systemto protect against dry runs. Theinitial temperature readingshould be recorded in the com-missioning data sheet.

SUMMARYThere are many options for test-

ing the performance and integrity ofcentrifugal pumps. The use of suchprocedures depends on the signifi-cance of the service and the natureof the pumpage. Users will find thatin most cases the standard compli-ment of manufacturing testing willbe sufficient. However, critical ser-vices involving serious environmen-tal or health risks may warrant theadded assurance of supplementaltesting. In either case, the user andmanufacturer must work as partnersto achieve the best engineering solu-tion for the particular application. ■

Leo Richard is a Technical ServiceManager with the Kontro Company,Inc. Mr. Richard has experience inprocess and project engineering withGeneral Electric and W.R. Grace.

An A range pump hooked up for testing.

Page 67: Centrifugal Pumps Handbook

68 The Pump Handbook Series

counter that internal clearances aredesigned to accommodate bearingwear, not to accommodate solid par-ticles. The crucial dimension, theysay, is bearing clearance and that isthe same for both types of seallesspumps. Further, canned motorpumps can effectively handle solidsif they are outfitted with externalflush or filters to remove particulatesfrom the pumpage before they circu-late around the bearings.

HIGH PRESSURE APPLICATIONSOne point on which all parties

agree is that magnetic drive pumpscannot tolerate as high pressures ascanned motors pumps can. A cannedmotor pump is a pressure vessel

erhaps you’ve decided to pur-chase sealless centrifugalpumps. The arguments arecompelling: zero emissions, no

need for complicated seal supportsystems, no need to replace expen-sive seals periodically. All manufac-turers of sealless centrifugal pumpsagree on these basic advantages. Buttheir agreement ends there.

As the two major camps in thesealless centrifugal debate — cannedmotor or magnetic drive — try toposition their chosen technology asthe most reasonable choice, they letloose a flurry of claims and counter-claims. It can get confusing.

To help you prepare for the bar-rage, we present advantages for bothtypes of sealless centrifugal pumps ascommonly stated by manufacturersand users. Consider the argumentsand decide which are most pertinentto your situation. Then you’ll be bet-ter prepared to discuss your specificconcerns with manufacturers.

THICKNESS OF CONTAINMENT SHELL Magnetic drive pumps can use

thicker containment shells since theirinner and outer magnetic rings do nothave to be as close together as therotor and stator in a canned motorpump. Manufacturers of mag-drivepumps claim that the thicker shell— up to five times thicker than thatof canned motor pumps — vastlyreduces the chances of breachingthe shell, especially as a result ofbearing wear.

Manufacturers of canned motorpumps counter that claim with twoarguments. First, the thicker shell ofa magnetic drive pump reduces oper-ating efficiency. Second, the shellmust be thicker (and the internalclearances wider) because magneticdrive pumps do not contain bearingmonitors. Unmonitored bearing wearcan cause the inner magnetic ring tocontact the shell. A thin shell wouldbe too prone to such damage. Cannedmotor pumps can use a thinner shellbecause bearings are closely moni-tored and bearing wear can be pro-jected from the data.

HIGH TEMPERATURE SERVICEPermanent magnets can toler-

ate heat better than motor wind-ings can. Thus, magnetic drivepumps can pump hot liquids — upto 750° F with just air cooling.Canned motor pumps can also beused in hot service but need watercooling jackets. Manufacturers ofcanned motor pumps agree thatcanned pumps should be watercooled for high temperatures. But,they reply, so should magneticdrive pumps since even rare earthpermanent magnets cannot tolerateextremely high temperatures.

SOLIDS HANDLINGWith greater internal clearances

and thicker containment shells, mag-netic drive pumps can handle solidsmore easily, their manufacturers say.Canned motor manufacturers

The Canned Motor vs.Magnetic Drive Debate

BY GREGORY ZIMMERMAN

P

CENTRIFUGAL PUMPS

HANDBOOK

Sealless canned motor pump designedfor hazardous liquids.

PHOTO COURTESY OF TEIKOKU USA

Page 68: Centrifugal Pumps Handbook

The Pump Handbook Series 69

since the stator windings lend addi-tional mechanical strength.

DIFFICULT-TO-HANDLE FLUIDSAccording to one manufacturer

of magnetic drive pumps, the biggestadvantage magnetic drives offer is theability to use non-metallics. Thesepumps are thus able to pump highlycorrosive materials, solvents, andother difficult fluids. That may betrue for some fluids, counter manu-facturers of canned motor pumps,but other issues are involved.Hazardous materials require failsafecontainment. Canned motor pumps,they point out, offer sealless doublecontainment. If the stator liningblows, a backup shell will contain thematerials. Doubly contained magnet-ic drive pumps rely on a mechanicalseal — the very thing we’re trying toavoid, say the manufacturers ofcanned motor pumps.

Canned motor proponents pointto another benefit of their technologyin hazardous environments: UL list-ing for the entire unit. Because acanned motor pump integrates theelectrical and mechanical portions,the entire pump must be UL listedfor use in, say, explosive atmos-pheres. Sundstrand canned motorpumps, for example, are tested undera procedure in which UL fills thepump with oxygen and ethylene andignites the gas. The explosion must becontained in the pump with no prop-agation of flames up or down the dis-charge piping. Magnetic drive pumpsare not UL listed — only the motorsneed to be. If a magnetic drive decou-

ples or loses a bearing, the skin tem-perature on the drive unit can exceedthe autoignition temperature of theexplosive compound.

COMPACT DESIGNCanned motor pump manufactur-

ers cite compact design as an addedadvantage. Canned motor pumps notonly save space but also require nofoundation work. Magnetic drive man-ufacturers counter that they can makecompact pumps by using a close cou-pled design. Besides, they add, theabsolute dimensions aren’t as impor-tant as meeting ANSI standard dimen-sions. ANSI standard dimensionsmake magnetic drive pumps easier toretrofit, according to their proponents.

ALIGNMENTCanned motor pumps have an

integrated single shaft and thus comeperfectly aligned from the factory.Alignment of the motor and magneticcoupling can be tricky in a magneticdrive pump.

A USER’S PERSPECTIVEFor one user, an engineer at a

chemical processing plant in the mid-west, UL area classification is themost important reason he preferscanned motor over magnetic drivepumps — in situations where cannedmotor pumps are optimal. This useralso relies on magnetic drive pumpsfor high temperature applications(e.g., heat transfer fluids), high horse-power requirements and for aqueoushydrochloric acid service (whichrequires nonmetallic pumps).

Another advantage this userstates for canned motor pumps is hecan predict bearing wear and thusschedule maintenance more easily.

FURTHER ADVICE ABOUT GOING SEALLESS

“We’ve been using cannedmotor pumps for some time and arecomfortable with the technology,”the user says. “Our electricians areadept at repairing the pumps — bothmechanically and electrically. Weare bringing in more magnetic drivepumps for certain applications and

we’re getting comfortable withthat technology, too,” he said.

This user and another engi-neer at a major chemical plantreport high reliability of bothtypes of pumps. Both report thatreliability increased as they gainedmore experience with seallesspumps. In each facility the majorcause of damage to sealless pumpsis operator and specification error.And as they learned to size thepumps correctly — to operate atthe best efficiency point of thepump — and to avoid operatingthe pumps off design, mean timebetween failure increased substan-tially. “We’ve gotten seven yearswithout failure from some of oursealless pumps,” said one user,“but we’ve also had cases wherewe replace the pump nearly everymonth because of dead head oper-ation, running dry or cavitation.”

All the above manufacturersagree that pumps must be speci-fied correctly for the applicationand that operators must be trainedadequately. “Users need to makesure we know everything aboutthe application,” says one manu-facturer. “We especially need toknow temperatures and vaporpressures at startup and shut-down, not just normal operatingconditions.”

Another key point: don’t sim-ply substitute pump problems forseal problems. In other words, ifyou’re faced with recurring sealfailures, be sure to root out thecause of the failure before yousimply bring in sealless pumps.Maybe the fault isn’t the seal. Ifthe problem lies elsewhere in thesystem, you’ll be left wonderingwhy your sealless pumps failedjust like the seals did. ■

Magnetic drive process pumpdesigned for zero leakage services.

PHOT

O CO

URTE

SY O

F GO

ULDS

PUM

PS, I

NC.

Page 69: Centrifugal Pumps Handbook

70 The Pump Handbook Series

What users

need to know

about the

National

Electric Code

and how

monitoring

options can

help.

ump users are no differentfrom other users of industrialprocessing equipment whomust comply with several

codes and government regulations. Itcan be a formidable task to keep upto date and appropriately apply rulesto specific situations. A greater effortis required to get code and regula-tions updated and clarified to keeppace with changing technology.Nonetheless, users need to under-stand the impact of the NationalElectrical Code on sealless centrifugalpumps and to know what monitoringoptions are available.

ELECTRICAL DEVICES IN HAZARDOUSLOCATIONS

The National Fire ProtectionAssociation (NFPA) has producedseveral codes for reducing the risk ofand damage from fires. One of theseis the National Electric Code (NEC).Section 500 of this document appliesto electrical devices operating in haz-ardous environments—where flam-mable/explosive materials are eitherroutinely or may be present in theatmosphere. These materials can begasses/vapors, liquids, a solid dust or

liquid mist. Since electrical devicesare present in these areas, the NECimposes requirements to reduce therisk of fires and explosions.

A Division 1 area is where explo-sive materials are routinely present inthe atmosphere (such as the bottom ofa spill containment, or a below-gradeinstallation where vapors could col-lect) and requires U.L.-approved elec-trical devices. Most sealless pumps,however, are operated in Division 2areas where explosive materials mayoccasionally be present in the atmos-phere. (In the chemical industry, thevast majority of materials are handledin closed systems.) The two areas arecovered in the NEC where they havethe potential to form an explosivecloud in the atmosphere. A growingcloud that comes in contact with asource of ignition, such as a hot elec-trical device, can cause a large explo-sion and fire.

To reduce this risk, the NECrequires that the Auto-IgnitionTemperature (AIT) be determined foreach stream in the process area aswell as its geographical area.Electrical devices intended for opera-tion in hazardous areas are also

BY: ROBERT MARTELLI

National Electric CodeImpact on SeallessCentrifugal Pumps

P

CENTRIFUGAL PUMPS

HANDBOOK

Page 70: Centrifugal Pumps Handbook

The Pump Handbook Series 71

required by the code to have “T rat-ings.” If users follow this section ofthe code, these electrical devices willnot constitute a potential source ofignition, vastly reducing the chanceof an explosion should an explosivecloud ever develop.

AITs are a concern for lighthydrocarbons including n-butane andacetylene, which have AITs below600°F; pentane and hexane, whichhave AITs below 500°F; and diethylether and heptane, which have AITsbelow 400°F.

NONELECTRICAL SOURCES OF IGNITION

Ignition by nonelectrical sources—for example, steam, heat transferlines and reactor vessel walls—arealso possible in process areas. TheNEC does not address these sources.Another NFPA code covers nonelec-trical sources of ignition in section30, the Flammable and CombustibleLiquids Code. Specifically, Chapter 5

requires that you take precautions toprevent the ignition of flammablevapors from nonelectrical sources.Preventive measures are to be deter-mined by an engineer and/or “theauthority having jurisdiction.” Since“hot surfaces” are normally presentin chemical processing environ-ments, one precaution typicallytaken is to handle materials in closedsystems.

EDDY CURRENT HEAT GENERATION

Canned motor and magneticdrive pumps with metallic liner/con-tainment shells generate heat due toeddy current loss. Eddy currents arecreated by changes in magnetic fieldstrength during pump operation in agiven area of a stator liner or contain-ment shell. In most applications,pumps will operate at temperatureswell below 400°F (which is belowmost AITs) because of the coolingeffects of the pumpage. In general,the eddy current heat source is not

regarded as heat produced by an elec-trical device and, therefore, not clear-ly addressed in the NEC.

CANNED MOTOR PUMPS INHAZARDOUS AREAS (DIVISION 2)

The “skin” or outside tempera-ture of the canned motor pump is anissue in hazardous areas. Thesepumps contain a thermal cut-outswitch, which is located in the statorwinding hotspot and shuts down themotor if its setpoint is exceeded. Theuser is required to wire this switchinto the motor control circuit. If themotor cooling is lost due to someupset or misoperation, the pump willheat up and eventually open theswitch and shut off the power, pre-venting an excessive “skin” tempera-ture on the can. If the pump is in avolatile liquid service, it’s usuallydestroyed. In most cases, the switchwill not protect the pump from dryrunning—it is there only to meet NECrequirements.

FIGURE 1

Mag drive pump cooling circuit flow temperature measurement is made after the fluid haspicked up eddy current heat and partial bearing heat.

Temperature monitor on containment shell by liquid exit from magnet area

Page 71: Centrifugal Pumps Handbook

72 The Pump Handbook Series

For canned motor pumps, theNEC currently covers only conduitseals. This is to prevent hazardouspumpage from traveling through theconduit system to the motor starterroom in case of a stator liner and pri-mary seal failures.

MAGNETIC DRIVE PUMPS (DIVISION 2 AREAS)

Mag drive pumps with metalliccontainment shells are not typicallyregarded as electrical devices, despiteeddy current generation. (Mag driveswith nonmetallic containment shellshave insignificant eddy current gener-ation and associated heat-up poten-tial.) These losses are about 17% ofthe maximum rated horsepower ofthe drive, which works out tobetween 2 and 3 KW of power lossfor a drive rated for 20 hp. If an upsetor misoperation results in dry run-ning, recent tests have shown that thecontainment shell temperature canreach 800°F to 1200°F in one to twominutes of continued operation.Furthermore, mag drive pumps thatoperate for several minutes with nocooling provided to the magnet areahave straw-blue rings in the areas ofthe strongest magnetic flux, indicat-ing temperatures of at least 900°F.Even more disturbing, with this typeof failure there is a good chance of aspill or release occurring!

Currently, there is no require-ment to monitor mag drive pumpsfor an abnormal condition andsubsequently shut down the pump. Inmost cases, containment shell tem-perature, motor power, or pump flowmonitoring with alarm and shut-down capabilities can greatly reducethe possibility of ever reaching unac-ceptable temperatures. Today thesemonitoring options are routinelyavailable. Most mag drive manufac-turers provide the option of contain-ment shell temperature monitoring.However, there is no widely acceptedagreement on the best monitoringmethod for mag drive pumps. Eachmethod has strong and weak points.

Several mag drive pump manufac-turers have recently taken steps to iso-late the containment shell from theoutside atmosphere, eliminating the aircooling used in some designs. This willnot necessarily prevent the migration

FIGURE 2

Mag drive pump temperature measurement is made on the coolingcircuit inlet. Temperature variations will be much smaller here.

of flammable vapors or gasses to comein contact with the containment shell.Depending on maintenance proce-dures, there is also a small possibilitythat the pumpage will leak undetectedpast an improperly installed gasket andcollect near the bottom of the contain-ment shell, next to the outer magnetassembly. Therefore make sure thecontainment shell temperatures do notrise above the AIT of the materials pre-sent. Some mag drive manufacturersoffer a leak monitor option for this sec-tion of the pump, partially addressingthis concern.

HEAT TRANSFER FLUID PUMPS

Mag drive and canned motorpumps with ceramic insulation onthe stator windings in heat transferservice may present a problem sincein some situations the suction fluidtemperature can be in the 600°Frange. If this temperature exceeds anAIT for other nearby materials, thereis no increase in safety by applyingthis portion of the NEC to eitherpump. In this case, the “skin” temper-ature already exceeds limitations setby the NEC for electrical devices!Current NEC interpretations by sev-eral users preclude the use of acanned motor pump with ceramic

insulation in these instances. Theonly option for a canned motor pumpis a unit with a cooling jacket, thenecessary service water lines, and aconventional stator with the appropri-ate thermal cut-out switch. The cur-rent mag drive claim, again, is thatthere is nothing in the pump thatmeets the definition of an electricaldevice; therefore, no special monitor-ing or shutdown devices are requiredby code. If the magnets can operateat the required service temperature,no cooling water is required.

Although difficult, it may be pru-dent to revise pump and piping lay-outs so that no low AIT materials arenear the pump.

CODE APPLICATION AND CLARIFICATION

A word of caution: The NEC hasbeen adopted by OSHA as a refer-ence standard and you are requiredto follow it as a minimum. Be carefulwhen interpreting the code andremember that common sense doesnot always apply! When makinginterpretations or determinationsregarding legal regulations, a teamapproach is advisable. More informeddeterminations are made, and mis-takes are less likely.

Temperature monitor on containment shell by liquidentrance to magnet area

Page 72: Centrifugal Pumps Handbook

If the NEC panel would clarifythe application of the code to magdrive pumps with metallic contain-ment shells, it would help pumpusers considerably. Specifically, doeseddy current generation fall withinthe code definition of an “electricaldevice”? And for electrical devices ofwhich the “skin” temperature alreadyexceeds an applicable AIT by non-electrical sources, does the NEC pre-vent its use?

A National Electric Code changecan occur no sooner than 1999, whenit’s scheduled for update. Until then,users will have to operate under thecurrent code, taking precautions asthey deem appropriate.

CONTINUOUS MONITORING: SAVING MORE THAN THE PUMP

Although monitoring adds cost,users can take advantage of automat-ic shutdowns for other abnormal con-ditions (such as dry running) beforethe pump is destroyed and providebetter assurance that AITs are notexceeded. Monitoring can alsoimprove pump reliability in handlingheat sensitive materials.

Monitoring is especially impor-tant with silicon/tungsten carbidebearings. (Most sealless pump manu-facturers offer carbide bearings atleast as an option.) Monitoring thesebearings requires a differentapproach than for carbon bearings inorder to extend life. Carbide bearings

can be more prone to sudden break-age and failure due to misoperation;hence, these conditions must beidentified and the pump automatical-ly shut down if encountered.

Pump monitoring is a relativelynew concept for most operationspeople and not well understood. Yetthese workers play a key role inimplementing monitoring methods.Everyone involved should havepatience in finalizing the alarm andshutdown setpoints for successfulimplementation of the methodsused. Nonetheless most users gothrough several “false shutdowns” oreven a pump failure before deter-mining the proper setpoints.

TEMPERATURE RISE MONITORING

For a single method, tempera-ture rise monitoring offers the bestoverall protection against most pumpfailures, including dead-head/verylow flow, dry run operation, andrestricted cooling circuit flow in themagnet area. Moderate cavitationand gas entrainment in the pumpageare also involved when they reachthe point of upsetting the cooling cir-cuit flow.

Two temperature points arerequired to implement this monitor-ing. One is on the containment shellof the pump. In all current designs,this point must be located betweenthe magnet assembly and the contain-ment shell flange limiting what tem-

The Pump Handbook Series 73

perature can be monitored. Note thedirection of the cooling circuit flownext to the temperature measurementpoint. A more sensitive measurementresults by monitoring at the exit pointfor the cooling circuit flow after it haspicked up heat from the magnet area(Figure 1). This flow configuration canfind this exit point near the rotatingmagnets in pumps that use a dis-charge-to-discharge pressure circula-tion with a pumping vane near therotating magnets. On pumps that use adischarge-to-suction pressure configu-ration to drive cooling circuit flow(that is, where the cooling circuit inletflows past the containment shell at themeasurement point, before the tem-perature rise takes place) temperaturerise monitoring will not be as effective(Figure 2). Canned motor pumps mayalso have temperature monitoringinstalled. More sensitive readings canbe taken when the monitor point islocated after the cooling fluid passesbetween the rotor and stator liners(Figure 3). With the temperature probein this location, dry run protection willnot be as effective as what can be pro-vided by power monitoring.

The second temperature moni-toring point is on the suction line orsupply vessel, providing suction tem-perature compensation and takes intoaccount temperature changes fromday to night, and seasonal variations.This greatly eliminates false shut-downs and failures. Keep enough dis-tance between this point and thepump to ensure that suction recircu-lation will not conduct heat from thepump and up the suction line to themeasurement point during dead-head operation. Locating this pointupstream of a suction basket strainermay provide enough isolation to beeffective. If you go to the tank for thistemperature, keep in mind that thesun can warm up the suction line andpump unit much faster than it canwarm the tank during nonoperation.If this happens and you get an inac-curate measurement, you may shutdown the pump on start-up whenthere is nothing wrong.

The temperature rise is deter-mined by the difference between thecontainment shell and suction tem-peratures (Figure 4). Pump supplierscan provide an expected “normal”

Canned motor pump cooling circuit temperature measurement madeat its hottest point.

Temperature moni-tor on rear bearinghousing at liquid exitfrom rotor-statorliners

FIGURE 3

Page 73: Centrifugal Pumps Handbook

74 The Pump Handbook Series

temperature rise. Typical alarm andshutdown points may be 10°C and20°C above this value. Field experi-ence will be required to finalize thesesetpoints for each application, sincethis is a relatively new concept forpump users. A good approach is tofind a temperature rise that is suffi-ciently far away from the normaloperating range (with its usual varia-tions), and that still results in liquid inthe containment shell, with a fewdegrees of boiling point margin left inthe magnet area. Pump suppliers canhelp by supplying pump cooling cir-cuit pressure. Knowing the pressure,you can calculate the liquid boilingpoint in the cooling circuit. The maxi-mum cooling circuit temperatureneeds to stay below this value.

Recent discussion about contain-ment shell temperature rises down-plays the effectiveness of containmentshell temperature monitoring. Somesay the temperature measurement isnot sensitive enough for the rapid risefound in dry running. However, tem-perature rise monitoring does notneed a large change in containmentshell temperature to be effective.

Also, temperature rise monitor-ing does not protect against motor orouter magnet bearing failure. Periodicvibration monitoring or additionalbearing temperature monitoring aretwo proven ways to protect againstthese types of failure.

Wear of the inner sleeve bear-ings may be detected by temperaturerise monitoring if there is enoughwear to alter the cooling circuit flowor the eddy current heat generation.This will depend on the pump usedsince temperature rise is not always adirect result of bearing wear.

FLOW AND LEVEL MONITORING

For processes where the supplytank level or pump flow are alreadymeasured and sent into a process con-

trol computer, adding these monitoringdevices can be relatively inexpensive.The only other hardware required isan output relay in the pump motorcontrol circuit. Then the software isprogrammed to implement a low levelshutdown or a high flow/low flowshutdown for the pump. The low levelmethod is effective against running thetank dry but does not cover othercommon pump failures.

Flow monitoring provides a littlebetter protection because it protectsagainst closed suction and dischargevalves in addition to dry running. Itcan also protect against excessive flow.The narrower the range between theshutdown setpoints and normal opera-tion, the better the protection; howev-er, false shutdowns must be avoided.Neither of these methods protectsagainst mild cavitation, a plugged cool-ing circuit flow path, or worn innerbearings. In services where theseother modes of failure are unlikely,this method can be quite effective.

POWER MONITORING

In this case, motor power (kilo-watt draw) is monitored, usually inthe pump starter room. It has theadvantage of not requiring anyprocess connections to install, makingit one of the easiest to incorporate intoexisting processes. This can eliminatecorrosion/erosion concerns in slurryor acid service where exotic materialsof construction are required. It is easi-er to establish shutdown setpoints ifthe pump is operating in the range of60% to 90% of its BEP.

As with any electronic device inan operating plant, it must not beaffected by radio frequency interfer-ence (e.g., portable radios). Powermonitoring has the same advantagesand disadvantages as flow monitoringin protecting against previouslydescribed failures.

Current monitoring of the motor

amp draw can also be effective aslong as the horsepower draw is nearthe motor’s nameplate rating.Otherwise, the amp draw versuspump curve becomes flatter, and it’smore difficult to determine realisticshutdown setpoints.

CONCLUSIONS

Until the National Electric Codeis either revised or clarified, userswill need to make their own determi-nations of the potential hazards ofpump operations and choose suitablemeans of reducing the resultingrisks. Efforts are underway at severalpump manufacturers to improvecontinuous pump monitoring. Amore universally accepted methodshould result. ■

Robert H. (Bob) Martelli isEngineering Specialist in FacilitiesEngineering, Dow Corning Corporation,Midland, MI.

Temperature rise is calculatedin a process control computerand compared to alarm andshut-down setpoints for appro-priate action.

FIGURE 4

TDI

1942

YS

1942TI

1942TI

1942

ALARM

SHUTDOWN

PG

PG

Page 74: Centrifugal Pumps Handbook

The Pump Handbook Series 75

ydrofluoric acid has touchedall of our lives because somany industries use it in theirmanufacturing processes. For

example, a beryllium-shafted golfclub and a coffee mug with an etcheddesign have been manufactured usinghydrofluoric acid. It has also beenused as a catalyst in the manufactureof ozone-friendly refrigerants. Yethydrofluoric acid is a potentially dan-gerous chemical. Acid leaks can yielddevastating effects, ranging fromtoxic fume inhalation to severe chem-ical burns, injuring people and dam-aging equipment.

Many plants pump hydrofluoricacid using traditional seal technology.Of course, mechanical seals can leak.Because of the potential dangerinvolved, hydrofluoric acid leaks arenot tolerated. One way to reduce thethreat of leakage is to use a seallesstechnology. Consider several key fac-tors when selecting a sealless pumpfor hydrofluoric acid applications,including proper metallurgies, com-patible bearing materials, andhydraulic and pump configurations.

METALLURGIES

Proper pump metallurgy is criti-cal for pumping hydrofluoric acid.Two primary variables dictate whatmetallurgies are necessary underpump operating conditions. The firstis temperature. Hydrofluoric acid issimilar to many other acids in that astemperature increases, so does theaggressive nature of the fluid. The

second variable is the percent con-centration. Table 1 and Figure 1define the most suitable metallurgiesfor given applications. When chemi-cal process industries use hydrofluo-ric acid, its nature is generallyaggressive. Consequently, worst casescenarios have more significance indecision-making choices.

Temperature and concentrationare not the only variables that impactcorrosion rates. Factors such asvelocity, aeration and other contami-nates play equally important roles inmetallurgical corrosion.

Five metallurgies (Table 1 andFigure 1) are suitable for any givencondition. Silver, goldand platinum are amongthe metals most resis-tant to hydrofluoric acidcorrosion. Two othermetallurgies are moreaffordable and provideexcellent results, main-taining corrosion at lessthan 20 mils per year(mpy) during adverseconditions. One is 66Ni32Cu (Monel 400), andthe other is 54Ni 15Cr16Mo (Hastelloy C-276®).There are some pitfallsin the composition of54Ni 15Cr 16Mo. Thisalloy is less resistant to corrosionthan 66Ni 32Cu, especially if oxygenis present; whereas 66Ni 32Cu isgenerally corrosive resistant, even totemperatures up to 300°F. (With

either alloy, stress-corrosion crackingmay be inevitable if water or oxygenare present, and in that case, corro-sion and cracking would be wide-spread and not localized.) Both ofthese metallurgies are excellentchoices for handling hydrofluoricacid.

BEARINGS

Bearing material is every bit ascrucial to a pump’s mechanical stabil-ity as its overall metallurgical compo-sition because the bearings areexposed to the acid. This is particu-larly important in canned motor tech-nology because the pumping process

is responsible for thecooling and lubricationof the bearings.

What is the properbearing material? Whatwill hold up under theunforgiving corrosive-ness of hydrofluoricacid? The answer is100% alpha grade sil-icon carbide, which is apressureless sintered sil-icon carbide. Bearingsmade of this materialcan withstand high tem-peratures and maintaindominating resistance tostrong acids. Alpha

grade has better resistance to wearand abrasion than the beta version ofsilicon carbide. However, both arepressureless sintered, or self-sin-tered, silicon carbide products.

BY: JOHN V. HERONEMA

H

Pumping Hydrofluoric AcidConsider proper metallurgies, compatible bearing materials, and hydraulic and

pump configurations when pumping this acid.

Temperatureand percent

concentrationdictate what

metallurgies arenecessary under

operatingconditions.

CENTRIFUGAL PUMPS

HANDBOOK

Page 75: Centrifugal Pumps Handbook

76 The Pump Handbook Series

Do not use reaction-bound sili-con carbides for hydrofluoric acidprocesses. These forms of silicon car-bide contain free silicon or graphitebecause reaction-bound silicon car-bides require silicon as a sinteringaid. Free silicon is subject to theattack of corrosive acids, resulting inbearing breakdown. In alpha andbeta grades of silicon carbide, no sin-tering aids are used, giving bothgrades almost complete chemicalinertness. The bottom line is thatthere is little difference between thealpha and beta grades of silicon car-bide. Most of the difference lies with-in the processing of the finalproducts. Nonetheless, alpha gradesilicon carbide is the preferred mater-ial for chemical processes that usehydrofluoric acid. Both alpha or betagrades of silicon carbide shouldexceed bearing expectations.

HYDRAULIC CONFIGURATIONS

For hydrofluoric acid applica-tions, the same challenges arise againand again: low NPSHA, low flow andhigh head. In centrifugal pumps, lowflow and high head yield low specific

speed, meaning poorhydraulic efficiency.

Specific speed is adimensionless number thatrelates the hydraulic perfor-mance of centrifugalpumps to the shape andphysical properties of itsimpeller. The equation tocalculate specific speed isshown in Figure 2. Wherelow flow and high head arerequirements, use a partialemission pump with anopen or closed radial vaneimpeller. A standard guide-line for pumping a fluid asvolatile as hydrofluoric acidis to keep the specific speedabove 200. This ensures thepump will maintain a rea-sonable hydraulic efficien-cy (25% to 30%).

Specific speed can beeasily manipulated byincreasing, gallon by gallon,the flow of a pump untilthe desired Ns value isachieved. Another way toimpact specific speed

includes increasing the rotativespeed. This technique is sometimesdifficult because many motors havefixed rotating speeds. To manipulatespeed, a variable frequency drivemust be used. A variable frequencydrive can increase the speed at whicha motor runs while maintaining aconstant voltage. However, thesedevices can be expensive. Regardless,the results are the same—increasedhydraulic efficiency.

Hydraulic efficiency is importantwhen pumping hydrofluoric acidbecause it has a steep vapor pressurecurve. Unproductive energy, which isa direct byproduct of inefficiency, islost in the form of heat. This addedheat must not be allowed to localizein the suction zone of the pump case.If it does, and suction pressure is notgreat enough to suppress vaporiza-tion, the pump may fail. The abilityto carry the heat away is directlyrelated to the specific heat of thefluid. Specific heat is the ratio of afluid’s thermal capacity to that ofwater at 15°C; in other words, afluid’s ability to carry away energy inthe form of heat. Unfortunately, thisthermodynamic property of hydroflu-

oric acid is fairly poor. If adequateNPSHA existed for most hydrofluoricacid applications, the ability of theprocess to dissipate heat would notbe crucial. However, NPSHA is oftenlacking.

NPSHA is the net pressure of aprocess fluid at the suction of apump. Having adequate NPSHA isimportant when pumping hydrofluo-ric acid because of the volatility ofthe process. The graph in Figure 3demonstrates the relationship of tem-

FIGURE 1

Zone definition for common metallurgies

Materials in shaded zone have repeatedcorrosion rate of <20 mpy

Zone 1 Zone 4

20Cr 30Ni 70Cu 30Ni25Cr 20Ni Steel 66Ni 32Cu1

70Cu 30Ni1 54Ni 15Cr 16Mo66Ni 32Cu1 Copper1

54Ni 15Cr 16Mo GoldCopper1 Lead1

Gold PlatinumLead1 SilverNickel1Nickel Cast Iron Zone 5Platinum 70Cu 30Ni1Silver 66Ni 32Cu1

54Ni 15Cr 16MoZone 2 Gold20Cr 30Ni Lead1

70Cu 30Ni1 Platinum54Ni 15Cr 16Mo Silver66Ni 32Cu1

Copper1 Zone 6Gold 66Ni 32Cu1

Lead1 54Ni 15Cr 16MoNickel1 GoldPlatinum PlatinumSilver Silver

Zone 3 Zone 720Cr 30Ni 66Ni 32Cu1

70Cu 30Ni 54Ni 15Cr 16Mo54Ni 15Cr 16Mo Carbon Steel66Ni 32Cu1 GoldCopper1 PlatinumGold SilverLead1

PlatinumSilver

1 = No air

TABLE 1. CODE FOR HYDROFLUORIC ACID GRAPH

250

225

200

175

150

125

100

75

C

121

93

66

38

10 20 30 40 50 60 70 80CONCENTRATION HF,%

BOILINGPOINT

2

3

4

5

6

71

TE

MP

ER

AT

UR

E

Page 76: Centrifugal Pumps Handbook

The Pump Handbook Series 77

perature to pressure. As the tempera-ture rises, the required pressure tomaintain the acid in a liquid phaseincreases, and the vapor pressurecurve becomes dramatically steeperat higher temperatures. Any point leftof the curve means the process is liq-uid; conversely, any point right of thecurve means the process is vapor. Ifhydrofluoric acid is being pumped at100°F, the NPSH must be equal to orgreater than 27 psi. If not, the processwill flash, resulting in a heavily cavi-tated or dry running pump.

To ensure adequate NPSHA, theheat input from the pump must beconsidered. Hydraulic temperature risecan be calculated. The equation (Figure 4) considers three variables:hydraulic efficiency, head, and the spe-cific heat value of a process. By usingthis equation and considering thevapor pressure versus temperature risecurve, you can predict if adequateNPSHA is provided. The following isan example.

PUMPING SPECIFICS

Fluid Pumped = HF acidHead (H) = 790 feetFlow (Q) = 20 gpmNPSHA = 7 feet, Mechanical NPSHR = 6 feetTemperature (P) = 95°FVapor Pressure (P.T. PSIA) = 25Specific Heat (BTU/lb°F) = 0.78Pump Hydraulic Efficiency (n) = 15%Specific Gravity (sp gr) = 0.92

1. Solve for the slope of the vaporpressure curve. Pick one temper-ature/PSIA point below thedesign temperature and another20°F above the operating point.Convert data into PSIA per °F.

17 PSIA – 36 PSIA = 19 PSIA 75°F – 115°F 40°F

= 0.475 PSIA/°F2. Solve for the maximum allow-

able temperature rise that canoccur before the HF flashes.

(NPSHA – NPSHR) sp gr = PSI 2.31 0.475 PSI per °F

= maximum allowable temperature rise °F(Actual)(7 – 6) 0.92 = 0.39 PSI

2.31 0.475 PSI per °F

= 0.83°F allowable temperature rise °F

3. Calculate hydraulic temperaturerise due to inefficiencies.

H (1 – n) = Temperature Rise778 x n x Cp

Wheren = hydraulic efficiencyCp = Specific heat

(Actual)790 (1 – 0.15) = 7.37°F

778 x 0.15 x 0.78

Conclusion:Slope of curve = 0.475 PSIA/°FMax allowable temperature riseallowed = 0.83°FTotal hydraulic temperature rise = 7.37°F

In this example adequateNPSHA is not being supplied. If thisdata is plotted on a vapor pressureversus temperature curve, the endresult is obvious—the HF is vapor(Figure 5).

To calculate how much NPSHAis necessary to keep the HF fromvaporizing:1. Total hydraulic temperature rise

= 7.37°F

2. Convert 7.37°F to PSIA usingcalculated vapor pressure curveslope and consider allowabletemperature rise (0.83°F)

7.37°F – 0.83°F (Allowable Temperature Rise) = 6.54°F6.54°F x 0.475 = 3.1 PSI

Convert 3.1 PSIA to feet2.31 x 3.1 PSIA

= 7.78 feetsp gr (0.92)

3. Thus, 7.78 feet in addition to cur-rent NPSHA must be provided.

Current = 7 feet+7.78 feet Newly Calculated = 14.78 feet Total NPSHR

These calculations are conserva-tive because they assume that thetotal temperature rise will take placeat the suction of the pump. Thistends to be valid at minimum flows,but is conservative at design flow.These examples do not encompassevery possible scenario that could beexperienced, but they are effective

Equation to calculate specific speed

Example A

Ns = NQ 1⁄2 Ns = 3550 (15 1⁄2)

H 3⁄4 900 3⁄4Ns = 83.7

Ns = Specific speedN = Revolutions per minuteQ = Capacity, at best efficiency, in gpmH = Total head developed by maximum

diameter impeller at best efficiency, in feet

( )FIGURE 2

140

120

100

80

60

40

20

0-50 0 50 100 150 200 250

PSI

DEGREES F

PSISeries 1

LIQUID STATE

STARTING POINTLESS HEAT INPUT FROM PUMP

VAPOR

AFTERHYDRAULICTEMPERATURERISE

FIGURE 3

Graph demonstrating the relationship of temperature to pressurefor hydrofluoric acid

Page 77: Centrifugal Pumps Handbook

78 The Pump Handbook Series

sure in the motor to keep the acid liq-uid. These configurations are oftenreferred to as pressurized or reversecirculation designs.

Magnetically coupled pumps arealso ideally suited to the handling ofhydrofluoric acids. Mag-drive pumpswith metallic containment shrouds(sometimes called cans) also produceeddy current losses that transmit heatto the pumpage. Some manufacturersoffer different internal circulationpaths, including rear-mountedimpellers to compensate for pressuredrop and temperature rise.

Several mag-drives are availablewith nonmetallic shrouds. In these noheat is produced due to eddy currentlosses, which increases overall pumpefficiency and decreases motorrequirements in most cases. Shroudmaterials include ceramics, siliconcarbide, PEEK and reinforced fluoro-plastics. Remember, factors such asinlet temperature, NPSH, contami-nates and system requirements mustbe taken into account with eithercanned motor or magnetically cou-pled pumps.

Regardless of the pump style,several auxiliary items can smooththe path toward safe and effectiveoperation. If low NPSHA is a factor,an inducer can lower a pump’sNPSHR. This can sometimes alleviatecostly system cha-nges. Over andunder current measuring relays effec-tively protect against dry running.These simple devices enable thepump operator to set a minimumamperage draw based on the specificfunctional curve amperage draw. If

the current dropsbelow the set point,the pump will auto-matically shutdown.

Thermowellsand temperatureswitches are alsoeffective in detect-ing overheating ofthe process withinthe pump. Oftenwhen pumps orsystems are experi-encing distur-bances, processt e m p e r a t u r e s

140

120

100

80

60

40

20

0-50 0 50 100 150 200 250

PSI

DEGREES F

PSISeries 1

LIQUID STATE

STARTING POINTLESS HEAT INPUT FROM PUMP

VAPOR

AFTERHYDRAULICTEMPERATURERISE

Vapor pressure versus temperature curve

FIGURE 5

guidelines in the determination ofadequate NPSHA. In addition tothese calculations, always multiplyyour final calculated NPSHR by a 1.3safety factor to ensure a successfulpump application. Clearly, increasingNPSHA can be an expensive proposi-tion. However, it may be lower incost than reinvesting money into aproblem pump caused by borderlineNPSHA versus NPSHR margins.NPSHA, flow and head play equallyimportant roles when selecting apump configuration.

PUMP CONFIGURATIONS

Several effective pump configu-rations exist for handling hydrofluoricacid. Before selecting a configuration,accurately evaluate the NPSHA ver-sus NPSHR, flow, head, efficiencies,and temperature rise.

Canned motor pumps offer twodesigns that are extremely effectivefor pumping hydrofluoric acid. Thesedesigns can pressurize the fluid in themotor to increase vapor pressuremargins or to reverse the motor flow(internal circulation) direction, rout-ing the heated process to the suctiontank rather than the pump. Thesedesign capabilities are important dueto the temperature gained from vis-cous drag, eddy current losses, andmotor inefficiencies. Although differ-ent, they fundamentally achieve thesame end result, keeping the processfrom flashing in the motor. The basicpremise of both designs is to increasethe pressure in the motor so that,even though the process temperatureis rising, there is still adequate pres-

increase. These devices signal a possi-ble problem and allow for review ofthe pump and system before extremedamage occurs. However, they arenot independent of process tempera-ture fluctuations and may be effectiveonly in constant temperature applica-tions. Bearing monitors are alsoimportant because they can detectproblems, such as fracturing with sili-con carbide bearing systems. Some ofthe most effective bearing wear mon-itors detect axial and radial wear.These monitors are important inscheduling proactive maintenanceversus reactive maintenance, whichis critical when unscheduled down-time can mean lost revenue.

If pump metallurgies, bearingmaterials, and hydraulic and pumpconfigurations are approached prop-erly, pumping hydrofluoric acid canbecome as routine as brushing yourteeth. ■

ACKNOWLEDGMENTSTable 1 and Figure 1 are © 1993

by Nace International. All rightsreserved by Nace. Reprinted by per-mission.

John V. Heronema has been withSundstrand Fluid Handling for six years.He has held positions in manufacturingand quality engineering, and is currentlya Product Engineer with SundyneCanned Motor Pumps, Arvada, CO.

H (1-n) = Hydraulic temperature rise due to778nCp inefficiencies of pump performance

H = Rated head in feet at design flown = Rated efficiency at design pointCp = Specific heat of a process fluid

defined as BTU/lb°F

FIGURE 4

Equation for hydraulic temperaturerise

Page 78: Centrifugal Pumps Handbook

The Pump Handbook Series 79

BY: MAURICE G. JACKSON

User Perspective: When toApply Mag Drive Pumps

Making the move for the right reasons.

ment or secondary control is speci-fied for your application.

CONSIDER LIFE CYCLE COSTSMagnetic drive pumps are often

the only alternative to meet govern-ment hazardous materials and safetyregulations, such as OSHA 1910,requiring stringent levels of contain-ment or control. In addition, manycompanies now have policies, odor-free imperatives for example, requir-ing strict control of emissions.However, for some zero emissionsapplications tandem seal pumps offera viable alternative to mag drives andlife cycle cost must be considered inthe selection criteria. Table 1 showscalculations of life cycle costs of tan-dem seal versus mag drive pumps fortwo specified applications. In the firstexample, the mag drive pump has asignificantly lower initial cost andoperating costs only slightly higherthan the tandem seal. In this applica-tion, because mag drives offer morereliable containment most userswould select the mag drive. In thesecond example, however, the magdrive proves to be more costly interms of both investment and operat-ing cost, and use of the mag drivecan not be justified in terms of costalone.

DON’T INSTALL MAG DRIVES TOOVERCOME SYSTEM PROBLEMS

Magnetic drive pumps shouldnot be installed to solve a mainte-nance problem, such as a trouble-some mechanical seal, without firstdetermining the real reason for theproblem. Once the problem has beenidentified, insure that installation ofmag drive pumps will not create aripple effect. Typical pump and sys-tem problems to watch for are:• cavitation

• operating too far from best effi-ciency point (BEP)

• Net positive suction head avail-able (NPSHA) too low

• slurries

• pump operating without liquidin the unit

COMMUNICATE WITH YOUR VENDORA user of magnetic drive pumps

should also be aware of potentialproblems and communicate withthe vendor to insure they are avoid-ed. For example, the drive motorshould always be sized smaller thanthe magnets to prevent decoupling ifthe impeller is overpowered.Decoupling of the magnets will cre-

agnetic drive centrifugalpumps offer an advantageover normal single mechani-cal seal centrifugal pumps

by preventing fugitive emissions fromleaking to the atmosphere. Givenproper application and operating pro-cedures, these pumps can perform foryears without failure. Rather thandiscuss the design of these pumps - asubject that has already been thor-oughly addressed in articles, papersand presentations - let’s review thejustification for installing mag drivesand provide installation keys to insurereliability of the investment.

KNOW SECONDARY CONTAINMENT ORCONTROL REQUIREMENTS

Secondary containment and sec-ondary control are important terms tounderstand when selecting your magdrive pump. Secondary contain-ment insures the fluid will be con-tained if the primary can fails. Somemag drive suppliers accomplish con-tainment by installing a secondarycan around the primary unit. If theprimary can develops a leak sec-ondary control insures the leakagewill be controlled to a definedamount, but not contained. In select-ing a mag drive pump be sure toknow whether secondary contain-

M

CENTRIFUGAL PUMPS

HANDBOOK

Page 79: Centrifugal Pumps Handbook

80 The Pump Handbook Series

ate excessive heat build-up in thefluid. Decoupling may also resultin a locked rotor due to failure ofthe pump bearings and can. Inaddition, because magnetic drivepumps are often used to pumplow – less than 1.0 – specific grav-ity fluids and are generally sizedfor such applications, an operatorshould be aware that employingthe pump for water or higher spe-cific gravity applications mayoverpower the motor or magnets.Failure to communicate fluidproperties may lead to additionalproblems. A vendor will need toknow more than fluid viscosityand specific gravity to size yourmagnetic drive pump. Usersshould also specify vapor pressurevs. temperature data and specificheat, as well as size and percent-age of solids for the fluid beingpumped.

USE PROTECTIVE INSTRUMENTATION TO INSURE RELIABILITY

To insure the reliability ofmag drive pumps, protectiveinstrumentation is recommended.Listed below are some typicalinstrumentation available and theirfeatures.

• Power meter – monitorspower to the motor drivingthe mag drive pump. Themeter can be used to preventdry and dead headed opera-tion. The power meter is prob-ably the best choice if you arelimited to the selection of onetype of monitoring instrumen-tation.

• Can thermocouple – mountedon the can, it senses dry run-ning and bearing problems.

• Bearing wear detector – isused to sense the position ofthe shaft or rotor. It can providean indication of the condition ofthe pump’s bearings.

• Level detector – is placed in thesuction or discharge piping toinsure liquid flow to the pump.

CONCLUSION

Recent developments in magnet-ic drive pumps, harboring manyfunctional and maintenance advan-tages for pump users, are testimonyto an exciting future for magneticdrive centrifugal pumps. ■

Maurice G. Jackson is aEngineering Associate in the EngineeringConstruction Division of TennesseeEastman Division of Eastman ChemicalCompany, Kingsport, Tennessee. He has25 years of experience in pump opera-tion, maintenance and engineering.

Life cycle cost calculations for Tandem Seal Vs. Magnetic Drive Pumps

TABLE 1

Based on one year1 operation

Application One: 400 gpm; 120 ft head

Pump Type Tandem Seal Magnetic Drive

Initial Cost $13800 $113502

Basic hp Required 17.5 20.4Seal hp Required 1.5 NATotal hp Required 19 20.4kW•h 129400 138900Power Cost $6470 $69453

Product Loss $300 $04

Maintenance Cost $600 $6005

Total operating cost per year $7370 $75456

Application Two: 100 gpm; 240 ft head

Initial Cost $7000 $80002

Basic hp Required 11.4 14Seal hp Required 1.5 NATotal hp Required 12.9 14kW•h 84426 95320Power Cost $4220 $47763

Product Loss $300 $04

Maintenance Cost $600 $6005

Total Operating Cost per year $5120 $53766

1 Table data based on operation 350 days per year, 24 hours per day. 2 Material of construction is 316 stainless steel.3 Electric power calculated at $50 per 1000 kW. Electric motor efficiency of 92%

assumed in calculation of kW usage. 4 Product loss calculated at $50 per pound.5 Maintenance costs assumed a failure once every three years. The failure modes are

assumed to be seal failure for the tandem seal and bearing failure for the mag drive. 6 Figures do not include initial cost.

Page 80: Centrifugal Pumps Handbook

The Pump Handbook Series 81

The causes of part failuresin sealless centrifugals maydetermine system andoperational problems.

ealless pump failures canhighlight system or opera-tional problems once takenfor granted or blamed on

mechanical seals. Once a pumphas failed, it should be taken apartto identify the broken part orproblem area. Frequently, a bro-ken part can indicate the cause offailure. By establishing and reme-dying the origin of the failure,pump service life can be extendedand future failures minimized.

The following describes partfailures and their causes that indi-cate system and operational prob-lems.

DAMAGED THRUST SURFACES(FRONT OR REAR)

Cause - operation below theacceptable minimum flow rate

Many pump users think ofminimum flow relative to temper-ature rise and bearing wear prob-lems. However, extreme low-flowoperation in a centrifugal pumpcan also create hydraulic imbal-ance of the impeller, generatingthrusting and vibration. Becausesealless pumps do not have theshaft overhang typical of sealedpumps, an imbalance can causeextreme axial shuttling of therotating assembly which maybreak thrust surfaces.

When a pump’s desired oper-ating point is at a very low flowrate, check with the pump manu-facturer for the minimum rate. Ifthe desired flow rate is below therecommended minimum, add arecirculation loop to increasethroughput and prevent hydraulicimbalance.

Cause - insufficient net positivesuction head (NPSH) available

A sealless pump may requiremore NPSH to insure that thehydraulic balance is maintainedand the bearing system containsenough fluid. NPSH problems can

result in shuttling of the rotating ele-ment, making it bang against thrustsurfaces, and this can lead to ruptureof the containment shell or liner.Repipe the system to reduce suctionpiping friction losses. Removingunnecessary valving or changing thepump elevation will solve the prob-lem.

Cause - water hammerWith sealless pumps, water

hammer can manifest itself by caus-ing failure of the thrust surfaces asthe rotating element is slammedagainst them. To solve the problem,review valve operating sequencesand piping arrangements. Slow downvalve closing speeds or change valvetypes to reduce water hammer.

INTERNAL SLEEVE BEARING FAILURECause - operating the pump dry

Sealless pumps require fluid tocool the bearings. Lack of fluid passingthrough the bearings causes thermalexpansion of the bearing or journal,depending on the particular design.This expansion constricts passages,increasing friction and heat, and there-by causing the pump rotating elementto lock up and cease operating.Alternatively, if the dry running opera-tion is short, the bearings may heat upenough to fracture due to thermalshock when fresh fluid is introduced.

The first solution is to avoid run-ning sealless pumps dry. If running

dry is possible due to the desiredmethod of operation, install arecirculation line around thepump to insure that fluid willalways be running through it.

Cause - low fluid vapor pressurePumping fluids at tempera-

tures close to their vapor pressurecan create problems. In a seallesspump, fluids close to vapor pres-sure can flash as the fluid, in pass-ing, picks up heat from thecontainment shell or bearings.This additional temperature risebrings liquids closer to their vaporpressure, and only a small amountof additional heat from the bear-ings may increase the liquid tem-perature above the vapor pressurelimits, causing the liquid to flashand preventing it from cooling thebearings. Bearing failure in a seal-less pump requires prompt atten-tion to minimize the cost of therepair and prevent external leak-age of the fluid.

If the fluid is close to thevapor pressure before it enters thepump, or if it has characteristicsthat suggest that a small tempera-ture change will produce a largechange in the vapor pressure, askyour sealless pump manufacturerto predict the expected tempera-ture rise in order to verify thatflashing will not occur in the bear-ings. To prevent flashing, somepump designs incorporate sec-

Magnetic Drive Pump

S

Interpreting Sealless Pump Failures

CENTRIFUGAL PUMPS

HANDBOOK

Page 81: Centrifugal Pumps Handbook

82 The Pump Handbook Series

ondary pumping devices toincrease pressure as fluid movesinto the bearings. Other designsoffer secondary cooling to solvethe problem.

CONTAINMENT SHELL FAILURE

Cause – drive magnet contact-ing the outer surface

When the antifriction bear-ings supporting the drive magnetin a magnetic drive pump fail, themagnet may contact and rupturethe containment shell. Such con-tact is indicated by grooves andrub marks on the external surfaceof the containment shell. If thepump design has safety rub ringsinstalled, check clearances toinsure they are correct. Replacethe rings when repairing theantifriction bearings.

Cause – internal pressure high-er than containment shelldesign limits

Water hammer can burst thecontainment shell. The suddenincrease in pressure can driverotating elements against thrustsurfaces and put increased shockinto the containment shell.Alternatively, the increased pres-sure alone can distort and burst

the containment shell. In this case,the containment shell will show signsof expansion or distortion from theinside out. Repipe the system toreduce suction piping friction loss.Removing unnecessary valving orchanging the pump elevation willsolve the problem.

Cause – pump hydrotest pressurewas above design limits

Caution should be used whenhydrotesting assembled pumps thathave nonmetallic containment shells.Shells using a fiber fill for strengthmay not rupture the first time theyare exposed to pressure, but thefibers inside the material may be bro-ken. If so, the next time pressure hitsthem, the shells may burst due toineffective fiber reinforcement.Review instruction manuals andtechnical data and do not hydrotestnonmetallic shells above recom-mended pressures.

Note - Hydraulic thrust balance

Magnetic drive pumps andcanned motor pumps frequentlyhave specific clearances thathydraulically control the amount ofthrust which the rotating elementsexperience. When thrust surfaces orbearings fail, the subsequent internalrubbing that takes place can increase

wear on balancing surfaces andreduce the effectiveness of thehydraulic thrust control. If onlythe failed part is replaced, the nextfailure may then result from lackof hydraulic thrust control. Whenin doubt, always replace ahydraulic thrust surface part thatis worn.

SUMMARY

The best solution to pumpfailures is always prevention.Pump products should be proper-ly applied at all times. Don’t hesi-tate to contact the manufactureror his representative to ask forhelp, and be sure to describe theapplication, installation and oper-ating conditions for the pumpthoroughly. Also, save all parts.An examination of them may pro-vide invaluable clues to the originsof pump failure, offering keys toovercoming systems or opera-tional defects. ■

ABOUT THE AUTHOR:Charles A. Myers, Director of

Sales and Marketing at IWAKIWALCHEM Corporation, Holliston,MA., has been working with seallesspumps for 14 years. He is active onANSI and API sealless pump’s stan-dards committees.

Canned Motor Pump

Rear ThrustSurface

Process Lubricated Sleeve Bearing

Process Lubricated Sleeve Bearing

Forward ThrustSurface

Stator Liner

Stator (Motor Windings)

Rotor Assembly

Page 82: Centrifugal Pumps Handbook

The Pump Handbook Series 83

BACKGROUNDElectric-motor-driven pumps

have been around for about 75 years,and so has the nagging problem ofthe shaft packing or seal. Becausewater was the common fluidpumped, it rarely became a danger-ous problem. However, as the chemi-cal industry developed, leakagebecame a major concern, and betterseals were needed and developed.

Industries are now under scruti-ny for hazardous emissions of alltypes, and must comply with cleanair and water regulations dictated byCongress and implemented by theEPA, OSHA, and other governmentagencies.

Currently, any leakage of liquidor gas is a problem and must be min-imized or eliminated. The state of theart for mechanical seals is in therange of 500 ppm leakage, with somereleasing as little as 100 ppm. Byusing secondary seals with drainageand control instrumentation, levelscloser to zero can be accomplished atincreased cost to the user.

THE BASIC PROBLEMIf we can accept that contacting

surfaces with relative motionbetween them will eventually wear,then we can conclude that in the caseof mechanical pump seals, leakagewill ultimately occur. So it is desir-able to do away with any shaft seal.By not penetrating the pump housingwith a shaft, the seal is eliminated,

Magnetic Couplings forSealless Pumps

Elimination of seals ends leakage concerns.

BY RONALD P. SMITH

along with any potential leakage.This can be done by totally enclosingthe impeller/pump assembly and iso-lating it from the prime mover.

The question is, how do youdrive the pump with no direct con-nection to a prime mover (motor)?

A SOLUTIONFortunately, we have a natural

force, magnetism, that can be used toour benefit. As children, we experi-enced the magnetic force of two mag-nets operating through a table top orpane of glass. One magnet would fol-low the other until the gap betweenthem became too large and reducedthe force. That basic idea is used insynchronous magnetic couplings.

There are two basic styles ofmagnetic couplings in use. Figure 1shows a face-face coupling andFigure 2 illustrates a co-axial design.

Magnetic couplings can be madeto develop almost unlimited forces,based on choice of material and scale.Coupling designs for hundreds of foot-pounds of torque are available.

One of the most fascinatingaspects of permanent magnet cou-plings is that although they exhibitpowerful forces of attraction andrepulsion, they require no outsidesources of power. If properly used,they last indefinitely.

COUPLING CHARACTERISTICSIn any synchronous coupling,

torque is developed in the same

way. The maximum attractive forcebetween the poles occurs when thepoles are aligned in opposite polari-ty. Maximum repulsive force occurswhen the same polarity is aligned.In both instances, the transverseforce (torque) is at a null (zero). Thelatter position is the least stable. Themaximum transverse force (torque)occurs between the two positionswhere the normal force is zero.Stable positions occur only once perpole pair, so in the case of a 10-polecoupling, there would be five stablepositions.

The proper application of a per-manent magnet coupling requiresknowledge of the maximum torqueproduced by the motor. This is typi-cally twice the amount produced atthe rated horsepower.

Running torque = (Rated Horsepower x 5,250) /rpm (ft-lbs)

In the case of a 5 Hp motor at1,800 rpm with no load, the runningspeed with about 3% slip is 1,750rpm and running torque is:

(5 Hp x 5,250) /1,750 = 15 ft-lbs.

However, the motor will developabout 30 ft-lbs peak torque duringline start, and a magnetic couplingmust have a peak torque rating atleast that high to prevent loss of cou-pling. Figure 3 displays the relation-ship of peak to running torque. Theamount of safety factor for the appli-

CENTRIFUGAL PUMPS

HANDBOOK

Page 83: Centrifugal Pumps Handbook

84 The Pump Handbook Series

cation will determine the exact designpoint on the curve. Slow start condi-tions can reduce the amount of peaktorque required in the coupling andprovide overload protection forimpellers in case of a mechanical jam.

In a coaxial coupling the radialforces are balanced if all of the mag-net segments are of equal strength.Concentricity of the inner and outerassemblies is also required for equalair gap distance. These factors devel-op “magnetic balance,” which is assignificant as physical balance inreducing noise and bearing wear.

Face–face couplings develop sig-nificant axial forces. When in thealigned, attractive mode, the force isat a maximum. At peak torque, theaxial force approaches zero. If slip-page occurs, it goes through a maxi-mum in the opposite direction. Thiscoupling design requires proper bi-directional thrust bearings on eachmember to handle the variable

forces. Inadequate bearings willallow air gap variations that causemechanical noise and can be selfdestructive. For this reason, theface–face coupling is generallyrestricted to special applications.

The torque of face–face cou-plings is limited by the allowablemaximum diameter of the assembly.A coaxial coupling can be madelonger for increased torque once amaximum diameter is reached. Thetorque is essentially linear with axiallength. This benefit makes the coaxi-al design the one of choice for mostapplications.

STIFFNESSIf rapidly fluctuating loads cause

mechanical resonance with thepump couplings, changing the num-ber of poles will modify the stiffness.Relationships of pole spacing to gaplength must be taken into account tomaintain design efficiency.

UNCOUPLING, SPECIAL CASESThe uncoupling phenomenon

limits torque and is very useful. Inpumps, it might protect the impellerfrom damage or detect unacceptablethermal conditions. Obviously, a cou-pling can be made to exceed thetorque of the motor, as in a mechani-cal coupling, and use motor thermalor electrical overload protection toshut down the system.

Most pumps are designed with acoupling that will not slip or uncou-ple within rated performance andproper motor application.

In the unique case where thepeak coupling torque is exceeded,slippage or uncoupling results. Theimpeller then stops, and no fluid ispumped. The seriousness of this sit-uation will depend on the applica-tion and coupling design. Thesystem should detect lack of flowand shut down the pump before anymajor damage occurs. A low-levelaudible warning may be heard fromthe coupling.

Inertia of the system will notallow “pick up” of the impeller mag-net until the motor is stopped. Beforerestarting the pump, the cause of theuncouple should be determined.

Running uncoupled for longperiods should be avoided. Becausethe impeller is not rotating and nofluid is being pumped, no fluid isbeing circulated through the con-tainment can and no cooling of thecoupling occurs.

When one magnet element isrotating past the other, a significantamount of energy is converted toheat, and because the inner unit usu-ally has a poor thermal escape path,it will get hot. If the temperaturerises past the design point, demagne-tization can occur. This is either tem-porary (recovered when the couplingcools down) or permanent (recoveredonly by remagnetizing).

If a coupling has a nonmetallicbarrier such as ceramic or plastic,there will be no additional uncou-pling effect.

Metallic barriers of stainlesssteel, Hastelloy, etc., will heat uprapidly due to eddy currents and,depending on the fluid contained,could represent a dangerous condi-tion. If the additional heat raises the

FIGURE 1

A face-face magnetic coupling

INPUT SHAFT

STEEL

MAGNET

NONMAGNETIC MATERIAL

MAGNET

STEEL

OUTPUT SHAFT

Page 84: Centrifugal Pumps Handbook

The Pump Handbook Series 85

temperature of the magnets, a furtherreduction of force by demagnetiza-tion is possible. Either of these casescould affect restart and necessitate a“cool down” before restart. Thismight not be a concern, becausesome time should be spent identify-ing the cause of the high torquerequirement.

CONTAINMENT BARRIERSThe containment barrier is a key

element to sealless pump success. Itprovides the primary fluid contain-ment and the “window” to coupletorque in the system. Like other ele-ments in the system, it usually is con-nected to the pump with a flange andan O-ring seal.

The containment barrier is also acritical part of any permanent magnetcoupling design. Its magnetic andelectrical characteristics affect theheating and power losses of the sys-tem. The wall thickness and associat-ed mechanical gaps determine themagnetic air gap and the amount ofmagnetic material required for a giventorque, and therefore significantlyimpacts the cost of the coupling.

Table 1 displays some of thecommon barrier materials, along withtheir benefits and related costs.

Permanent magnet couplings caneasily handle larger air gaps than

allowed in canned motors. This is amajor benefit when handling highviscosity fluids or when suspendedparticles are in the fluid. Magneticparticles are to be avoided becausethey may collect between the magnetand barrier. Air gap clearance oneither side of the barrier should be assmall as possible, but their sizedepends on allowable bearing wear.If either rotating magnet assembly isallowed to contact the barrier, thepressure vessel may be compromisedand failure can occur. Mechanical“rub rings” or proximity detectors canbe used to indicate bearing failure.

It is usually not practical to makethe coupling barrier shell an integralpart of the pump housing. This shellshould have the thinnest wall possi-ble that satisfies the design pressurerequirements. The material must benonmagnetic and preferably non-

metallic to reduce eddy current heat-ing and associated power losses.

There are many barrier designs,the most common being plastic shellsfor small pumps and stainless steelfor large pumps, with pressurerequirements up to thousands ofpounds per square inch. Chemicalsbeing pumped dictate the choice ofmaterial, and frequently the shell ismade of the same material as thepump housing.

Ceramic shells, coated metal,and laminated metal are used forspecial cases.

In the case of solid metallic barri-ers, eddy current heating is devel-oped. This is torque transfer loss andcan amount to 5–10% of the inputpower. Generally ignored in smallsystems, it may be significant withmotors over 100 Hp. Most coolingcan be through the fluid if a generousflow within the barrier is established.Additional heat dissipation throughthe pump housing is possible if thebarrier shell has a metal–metal con-tact to the housing. Eddy currentheating can be reduced by loweringthe speed of the motor, as losses areproportional to speed.

Because the containment barrierbecomes a pressure vessel, fabricationtechniques are important. Designs areguided by ASME standard and manu-facturing processes dictated by quanti-ty. Small- to medium-size barriers areusually machined from solid bar stockin small quantities. Spun, hydro-formed, or deep-drawn shells may bemore economical in large quantities.Welded units, which are feasible in allsizes, require close process control toavoid stress corrosion problems.Pressure testing may be part of partcertification.

FIGURE 2

A co-axial magnetic coupling

Material Wall Pressure Chemical Eddy Current Relative Thickness Capability Resistance Heating Cost

Plastic Medium Medium High No LowCeramic Thick Low/Medium Medium No HighStainless Steel Thin High High Yes MediumHastelloy Thin High High Yes HighTitanium Thin High High Yes High

TABLE 1- BARRIER MATERIAL COMPARISON

Barrier(Flange sealto pump)

Pump Shaft

Driver

MotorShaft

MagnetsFollowerAssembly

Page 85: Centrifugal Pumps Handbook

86 The Pump Handbook Series

COUPLING DESIGNIn concentric couplings, the dri-

ving element connected to the motoris usually the outer magnet assembly.This part of the magnetic elementshas the highest mass and inertia. Itcan be made of magnetic iron andpainted, plated, or coated as neces-sary. It is usually not subjected tohigh corrosion environments anddoes not require special sheathing.

Salient magnet poles have gapsbetween them. These may be filledwith epoxy or other potting com-pounds to improve cleaning and min-imize magnet damage duringassembly.

The inner magnet assembly isthe “follower.” Because it is in theprocess fluid, special care must betaken to prevent corrosion or contam-ination of the pumpage. This assem-bly typically has rare earth magnetsmounted on an iron ring. If these

parts will be corroded by the fluid,they are sheathed or coated withappropriate materials. Frequently, astainless hub is used with a stainlesssheath welded to it, totally encapsu-lating the magnet assembly. Welding

techniques are critical to long-termlife, and leak or pressure testing isadvised.

Bathed in the process fluid, thefollower assembly may be affected bytemperature extremes. For tempera-tures over 100°C, efficienciesdecrease and designs become morematerial specific.Factors affecting design are:

• magnetic gap length (barrier +clearances)

• peak torque required

• space available for coupling

• form factor desired for comple-ment to pump (diameter x length)

• stiffness required

• fluid and corrosion concern

• maximum operating temperature

• running speed

• shaft sizes

• barrier material and allowableeddy current heating

INDUSTRY STANDARDSPermanent magnet couplings are

an integral part of a pump. Due pri-marily to the need for critical align-ment, they are sold as a unit with themotor. Pumps are manufactured tomeet industry standards such asthose published by ANSI, API, andthe Hydraulic Institute. These organi-zations have recently included stan-dards for magnetic couplings.

Standards are based on voluntarycompliance and in most cases insureinterchangeability of parts amongmanufacturers. To meet customerneeds for quality in specific applica-tions, pump manufacturers have theirown rigorous standards.

Permanent magnet materials arealso produced to industry standardsthat allow sizeable variation in mag-netic properties within materialgrades. Critical applications requiregreater control of properties and/orselection of parts for uniformity.

MAINTENANCEBecause there is no mechanical

wear in a magnetic coupling, thereshould be no need for maintenance.However, bearings do wear, andoccasionally a pump will require dis-assembly or a motor will need to bereplaced. The manufacturer willhave recommendations for handlingthis operation. Disassembly of themagnetic coupling should be doneonly by trained personnel using theproper fixtures.

Permanent magnet couplingscontain some of the most powerfulmagnet materials ever made, andthey are always energized. In sys-tems of 5 Hp and over, the forcesare greater than a person can controlby hands alone. Large pump cou-plings have axial forces in the hun-dreds of pounds. Fixtures ormechanical means to guide the partsand prevent damage to the contain-ment barrier are required during dis-assembly and re-assembly. Trainingis also required to avoid personalinjury. Magnets can be very unfor-giving of mistakes. ■

Ronald P. Smith is Manager ofEngineering for the Magnetic MaterialsDivision of Dexter Corporation.

Coupling Max.

Motor Max.

Running/OperatingTorque

To

rqu

e

Torque vs. Angle of Rotation

Angle of Displacement

FIGURE 3

The relationship of peak to running torque

Page 86: Centrifugal Pumps Handbook

The Pump Handbook Series 87

Suction Side Problems - Gas Entrainment

BY: JAMES H. INGRAM

ultiple symptoms associatedwith noncondensable suc-tion side gas entrainment,such as loss of pump head,

noisy operation, and erratic perfor-mance, often mislead the pump opera-tor. As a result, entrained gas isgenerally diagnosed by eliminatingother possible sources of performanceproblems. To adequately control gasentrainment a user should first beaware of systems most likely to pro-duce gas, and then employ methodsor designs to eliminate entrainmentinto these pumping systems.

ENTRAINMENT VERSUS CAVITATION The audible pump noise from

noncondensable entrained gas willproduce a crackling similar to cavita-tion or impeller recirculation.However, cavitation is produced by avapor phase of the liquid which iscondensable, while noncondensableentrained gas must enter and exit thepump with the liquid stream.

To test for gas entrainment overmild cavitation, run the pump backupon the curve by slowly closing thedischarge valve. The noise will dimin-ish if it originated from cavitation andthe pump is not prone to suction recir-culation. In contrast, with entrainedgas, continued performance at thisportion of the curve will choke off orgas-bind the pump, causing unusuallyquiet operation or low flow.

M GAS BOUND IMPELLERSAs a process stream containing

entrained gas nears the impeller, theliquid pre-rotating from the impellertends to centrifuge the gas from theprocess stream. Gas not passing intothe impeller accumu-lates near the impellereye. As entrained gasflow continues to in-crease, the accumulat-ing groups of bubblesare pulled through theimpeller into the dis-charge vane area wherethey initiate a fall inflow performance. Thebubble choking effect atthe impeller eye pro-duces a further reduc-tion of Net PositiveSuction Head Available(NPSHA). At this stagelong term damage to thepump from handlingentrained gases is gener-ally negligible whencompared with thedamage due to cavita-tion. If the processstream gas volumeincreases, however, fur-ther bubble build-upwill occur, blocking offthe impeller eye andstopping flow (Ref. 1).

A pump in this gas bound state,will not re-prime itself, and the gas,with some portion of the liquid, mustbe vented for a restart against a dis-charge head. The effort to restart agas bound impeller depends on

Capacity in U.S. Gallons per Minute

Head

in F

eet

Size: no. 55; Type: SQ. Speed: 1750 Impeller Diameter: 11”

Air quantities given are in terms of free air at atmosphericpressure referred to % of total volume of fluid being handled.

FIGURE 1. ENCLOSED IMPELLER-ENTRAINEDGAS HANDLING PERFORMANCE

NO AIR HEAD2%

5%

8%

10%

12%15%

12001000800600400200050

60

80

100

120

140

160

The LaBour Company, Inc. Effect on head and capacity ofvarying quantities of air with water being pumped.

CENTRIFUGAL PUMPS

HANDBOOK

Page 87: Centrifugal Pumps Handbook

head loss at 5% gas volume, theGould’s open impeller experiences a12% head loss at this volume. Someopen impeller paper stock designscan actually handleup to 10% entrainedgas because clear-ance between thecase and impellervanes allows moreturbulence in theprocess fluid, whichtends to break upgas accumulationmore efficientlythan an enclosedimpeller with wearrings. In addition,other designs, suchas a recessed im-peller pump, mayhandle up to 18%entrained gas. Infact, most standardcentrifugal pumpshandle up to 3%entrained gas vol-ume at suction con-ditions withoutdifficulty. (Ref. 2

discusses open impeller pump modi-fications.)

SYSTEMS PRODUCING ENTRAINED GAS

The most common conditions ormechanisms for introducing gas intothe suction line are: 1. Vortexing

2. Previously flashed process liquid conveying flashed gasinto the suction piping.

3. Injection of gas, which does notgo into solution, into thepumpage.

4. Vacuum systems, valves, seals,flanges, or other equipment in asuction lift application allowingair to leak into the pumpagestream.

5. Gas evolution from an incom-plete or gas producing chemicalreaction.

If a particular application pro-duces entrained gas or has the poten-tial to do so, the best solution is toeliminate as much entrainment aspossible by applying corrective pumpsystem design and/or a gas handlingpump. If liquid gas mixing is desired,employ a static mixer on the dis-

impeller position, type and valvingarrangement, among other variables.Degassing is easier to accomplishwith a variable speed driver, such asa steam turbine, than with a constantspeed electric motor drive. In addi-tion, a recycle line to the suction ves-sel vapor space is often an effectivemethod for degassing an impeller,since with this arrangement thepump is not required to work againsta discharge head. (Ref. 1 describesmethods for venting gas on modifiedpumps that are gas bound.)

As a rule, if the probability ofentrained gas exists from a chemicalreaction, the inlet piping designshould incorporate a means to ventthe vapor back to the suction vessel’svapor space or to some other source.

EFFECTS OF ENTRAINED GAS ONPUMP PERFORMANCE

Figures 1 and 2 illustrate theeffect of entrained gas on a LaBourenclosed impeller and a Gould’spaper stock open impeller. As illus-trated by the figures, 2% entrainedgas does not produce a significanthead curve drop. Note that while theLaBour impeller experiences a 22%

Gallons per Minute0 500 1000 1500 2000 2500 3000

FIGURE 2. OPEN IMPELLER-ENTRAINED GAS HANDLING PERFORMANCE

Gould’s Pumps, Inc. Approximate Characteristic Curves of Centrifugal Pump

4%

0%2%

6%0%2%4%6%

6%0%Bhp@ sp gr=1.0

Size: 6x8-18 Speed: 1780 rpm Impeller Diameter: 17 1/4”

50

60

70

80

Effic

ienc

y %

250

200

150Brak

e Ho

rse

Pow

er (B

hp)

150

200

250

300

350

Head

in F

eet

FIGURE 3. DEVELOPMENT OF A VORTEX

A. Incoherent surface swirl

B. Surface dimple with coherentsurface swirl

C. Vortex pulling air bubbles tointake

D. Fully developed vortex with air core to nozzle outlet

AIR

(c)

(a)

(b)

(d)

88 The Pump Handbook Series

Page 88: Centrifugal Pumps Handbook

The Pump Handbook Series 89

charge of the pump. In addition, ananticipated drop in pump head due toan entrained gas situation may be off-set by oversizing the impeller.

Of the five aforementionedmechanisms, vortexing is the mostcommon source of entrained gas.Therefore, a user should be especiallycautious employing mechanicalequipment, such as tangential flashgas separators and column bottomsre-boilers, likely to produce a strongvortex.

VORTEX BREAKER DESIGNThe extent of gas entrainment in

the pumped fluid as the result ofvortex formation depends on thestrength of the vortex, the submer-gence to pump suction outlet, andthe liquid velocity in the pump suc-tion nozzle outlet. Vortices form notonly through gravity draining vesselapplications, but also in steady statedraining vessels, and in vesselsunder pressure or with submergedpump suction inlets. Vortex forma-tion follows conservation of angularmomentum. As fluid moves towardthe vessel outlet, the tangentialvelocity component in the fluidincreases as the radius from the out-let decreases. Figure 3 shows variousstages of vortex development. Thefirst phase is a surface dimple. Thisdimple must sense a high enoughexit velocity to extend from the sur-face and form a vortex. (For experi-mental observations regardingvortex formation see Refs. 3, 4.)

The most effective method to

eliminate entrained gas in pump suction piping is to prevent vortexformation either by avoiding vortexintroducing mechanisms or by em-ploying an appropriate vortex break-er at the vessel outlet. A ”hat” typevortex breaker, illustrated in Figure4, covers the vessel outlet nozzle toreduce the effective outlet velocity.This design doesn’t allow a vortex tostabilize because the fluid surfacesenses only the annular velocity atthe hat outside diameter (OD). Inaddition, the vanes supporting thehat introduce a shear in the vicinityof the outlet to further inhibit vortexformation. An annular velocity of1/2 ft/sec at the hat OD produces aviable solution. Variations in hatdiameters from 4d to 5d and hatannular openings of d/2 to d/3 areacceptable when annular velocity cri-teria are met. Annular design veloci-ties of more than 1 ft/sec are notrecommended.

”Cross” type breakers, installedabove or inserted in vessel nozzle out-lets as shown in Figure 5, work forsome applications by providing addi-tional shear to inhibit a mild vortexfrom feeding gas into a nozzle outlet(providing enough submergence isavailable). However, this design willnot stop a strong vortex and willdecrease NPSHA. A user should beaware of these limitations.

COLUMN VORTEXINGIf a column draw-off pump is

erratic and/or nearly uncontrollable, avortex may be feeding gas into the

draw-off nozzle of the pump as illus-trated by Figure 6a.

It may be difficult to understandhow a pump with 60 ft of verticalsuction could be affected by en-trained gas, but in this real caseexample Murphy’s law applied twice.First, since the pump system in ques-tion has a NPSHA greater than 50 ft,the piping designer employed a small-er suction pipe with a liquid velocityof 10 ft/sec. Second, the columndraw-off nozzle was sized accordingto normal fluid velocity practice. As aresult, the tray liquid had an exitvelocity of 5 ft/sec with a liquid level6-in. above the top of the draw-offnozzle and a vortex formed, feedinggas into the draw-off nozzle.

As in the above example, due to alack of proper submergence, gas is car-ried into the pump suction piping as ahigh liquid downward velocity exceedsthe upward velocity of a gas bubble.

Many draw-off vortexing prob-lems may be eliminated by properpump system design or by one of twovortex breaker designs illustrated byFigures 6b and c. The selection of thebreaker design may depend on thedowncomer arrangement and spacelimitations. The most effective vortexbreaker is the slotted pipe designshown in Figure 6c.

Application of these correctivepump systems designs or installa-tion of an appropriate gas handlingpump can solve suction side gasentrainment problems, resulting in asmoother process operation. ■

FIGURE 4. “HAT” TYPE VORTEX BREAKER FIGURE 5. “CROSS” TYPE VORTEX BREAKER

Page 89: Centrifugal Pumps Handbook

90 The Pump Handbook Series

REFERENCES1. Doolin, John H., ”Centrifugal

Pumps and Entrained-Air Problem,”Chemical Engineering, pp.103-106(1963)

2. Cappellino, C.A., Roll, R. andWilson, George, ”CentrifugalPump Design Considerations andApplication Guidelines forPumping Liquids with EntrainedGas,” 9th Texas A&M PumpSymposium 1992

3. Patterson, F.M., ”Vortexing canbe Prevented in Process Vesselsand Tanks,” Oil and Gas Journal,pp. 118-120 (1969)

4. Patterson, F.M., and Springer, E.K.,”Experimental Investigation ofCritical Submergence for Vortexingin a Vertical Cylinder Tank,”ASME Paper 69-FE-49 (1969)

5. Kern, Robert, ”How to DesignPiping for Pump Suction Con-ditions,” Chemical Engineering,pp.119-126 (1975)

James H. Ingram is an EngineeringTechnologist with Sterling Chemicals inTexas City.

Figure 6a. Tray take off nozzle with vortex from lack of correctsubmergence and too high exit velocity.

10”6”

BUBBLE CAP

10’/sec.

DOWNSPOUT ORDOWNCOMER FROMTRAY ABOVE

Figure 6b. Plate extension over outlet nozzle lowers high outletvelocity.

1/2’/sec.

EXTEND PLATE FROMVESSEL WALL. CHECK VELOCITY AT PLATEEDGE ≤ 1/2’/sec.

Figure 6c. Slotted pipe vortex breaker.

AREA OF SLOTS—3XPIPE CROSS SECTIONAREA. CHECK VELOCITY INTO SLOTAREA ≤ 1’/sec.

FIGURE 6. DESIGN MODIFICATIONS FOR A SYSTEM EXHIBITING A LACK OFADEQUATE SUBMERGENCE AND PROHIBITIVELY HIGH EXIT VELOCITY

Page 90: Centrifugal Pumps Handbook

The Pump Handbook Series 91

his past year’s Texas A&MInternational Pump UsersSymposium at the George R.Brown convention center in

Houston, TX included a discussiongroup entitled Nozzle Loading andPump Operability co-coordinated byJohn Joseph of Amoco Oil and WillieEickmann of Houston Lighting andPower. According to Gary Glidden,also a discussion leader for thisgroup, the two day discussion was a”standing room only” affair. Clearly,nozzle loading is a subject of concernto pump users.

ESTABLISHED LOADING STANDARDS QUESTIONED

Much of the discussion focusedon the difficulty of establishing stan-dards for allowable nozzle loads.Although the current API 610 7th edi-tion standard for centrifugal pumpsin general refinery service providesvalues for maximum loads (Table 1,Figure 1), many pump users believethe API allowable loads are toohigh—especially for use as specifica-tions for installation designs whichfail to recognize the possibility of”unplanned” stresses on the piping,such as those produced by founda-tion settling. However, as noted byJames E. Steiger in his paper, API 610Baseplate and Nozzle Loading Criteria,”Before the 6th Edition of API 610was published, there were no indus-

BY: KIMBERLY FORTIER, ASSISTANT EDITOR

T

Nozzle Loading –Who Sets the Standards?

Or, to what extent should the pump be used as a piping anchor?

try accepted standards for allowablepiping loads acting on centrifugalpumps.” Moreover, when these pip-ing load standards are absent or notuniversally accepted by pump users,

manufacturers and piping engineers,these groups tend to set independent,often contrary standards, furthercomplicating the design process.

In an attempt to overcome these

Note: Each value shown below indicates a range from minus that value to plus that value; for example, 160 indicates a range from -160 to +160.

Nominal Size of Nozzle Flange (inches)

Force/Moment* 2 3 4 6 8 10 12 14 16

Each top nozzleFX 160 240 320 560 850 1200 1500 1600 1900 FY 200 300 400 700 1100 1500 1800 2000 2300FZ 130 200 260 460 700 1000 1200 1300 1500FR 290 430 570 1010 1560 2200 2600 2900 3300

Each side nozzleFX 160 240 320 560 850 1200 1500 1600 1900FY 130 200 260 460 700 1000 1200 1300 1500FZ 200 300 400 700 1100 1500 1800 2000 2300FR 290 430 570 1010 1560 2200 2600 2900 3300

Each end nozzleFX 200 300 400 700 1100 1500 1800 2000 2300FY 130 200 260 460 700 1000 1200 1300 1500FZ 160 240 320 560 850 1200 1500 1600 1900FR 290 430 570 1010 1560 2200 2600 2900 3300

Each nozzleMX 340 700 980 1700 2600 3700 4500 4700 5400MY 260 530 740 1300 1900 2800 3400 3500 4000MZ 170 350 500 870 1300 1800 2200 2300 2700MR 460 950 1330 2310 3500 5000 6100 6300 7200

*F = force, in pounds; M = moment, in foot-pounds; R = resultant

TABLE 1. API ALLOWABLE NOZZLE LOADS

CENTRIFUGAL PUMPS

HANDBOOK

Page 91: Centrifugal Pumps Handbook

92 The Pump Handbook Series

complications, one user developed astandard operating procedure basedon measured changes in pump align-ment to be applied universallythroughout their plant. Changes inalignment subsequent to connectionof suction and discharge lines indi-cate shaft deflection. This user set themaximum shaft deflection at 0.002”regardless of pump size or configura-tion. However, because this proce-dure relies on establishing a baselinealignment before the lines are con-nected, this standard cannot beapplied to all pumps. For example,the feedwater pumps Glidden oper-ates at Houston Lighting and Poweremploy welded nozzles, which don’tallow the pump to be isolated inorder to determine the ”zero-load”alignment, as opposed to flanged nozzles. And, since this proceduredepends on measuring alignmentrather than forces and moments asfor the API specifications, making acorrelation between the two stan-dards is ”almost apples and oranges,”says Glidden.

Figure 2 illustrates a commonconsequence of nozzle loading on apump. While the case bows in onedirection due to piping loads, theshaft sags in the opposite direction asa result of thermal deformation.Pump operators witness the endresults of overloading a pump nozzlein misalignment, vibration, bearing

or coupling failure, and shortenedseal life; but, according to some, theestablished standard fails to addressthe correlation between loading lev-els and these failure modes. And,these users are concerned that rela-tively slight levels of nozzle loading,even those within API specifications,may have costly ramifications, interms of downtime and pump life, inthe long run. In fact, according toJoseph, the discussion at the PumpUsers Symposium quickly pro-gressed beyond the question, ”Howmuch (loading) is too much,” towhether the pump should ”even beconsidered an anchor for the piping.”Joseph concludes, ”The piping(should) exert as little force as practi-cally possible during operation.”

CAREFUL PIPING DESIGN CAN REDUCE STRESSES

But how much is ”practicallypossible”? Joseph recommends”shooting for 10% or less of API(allowable nozzle loading specifica-tions) during running conditions.” Healso points out, however, that pipingstresses can be reduced to zero, ”Mypersonal preference under hot condi-tions is that the piping exert nearlyzero forces and moments.” In mostcases to obtain zero stress under hotconditions requires exerting somestress on the piping in the cold condi-tion. These stresses will then relax

with the thermal growth that occursduring hot operation. Joseph favorscareful calculations in the designphase to insure that ”the spring hang-er forces and the deflection of thebeam they’re supported from match-es the weight and growth of the pip-ing when it’s full of liquid attemperature.” For example, a springhanger supporting a 20’ straight verti-cal piping section might relax a full1/4-1/2” due entirely to thermalgrowth in the vertical direction. Addthis growth to the pull of the process-liquid weight and the result is, thepiping stress and strains during hotrunning differ drastically from thoseprior to start-up. One operator actual-ly measured a 0.150” horizontalmovement of the pump. ”You’ve gotto think, what is it (the pump) goingto look like with a hot flow of liquidand then back calculate to the cold,empty position you want that pipeat,” says Joseph. To obtain minimumloading during the running condition,the pipe should be supported in aposition requiring it to be pulleddown to the pump. During hot opera-tion, the thermal growth of the pipingand the weight of the liquid will thendepress the piping into the relaxedposition.

ECONOMICS

Piping engineers counter thesearguments for low to zero pipingloads, claiming, as Steiger notes, that”the pump manufacturers and rotat-ing equipment engineers are too con-servative and the higher piping loadsdo not usually lead to significantoperability problems.” The largerpiping loads are desirable becausethey result in simpler and signifi-cantly less expensive piping configu-rations. Yet, the 1985 PressureVessel Research Committee (PVRC)Pump-Piping Interaction ExperienceSurvey indicates that there is ”a significant pump-piping interactionproblem and that it has an annualimpact on the order of one half- billion dollars” (Ref. 1). ”Economicsplays a big role in these decisions,”adds Glidden, ”However, if you do abad job up front, this will com-pound, resulting in a terrible-run-ning pump.”

Shaft Centerline

X

X

Y

Y Y

Y

X

X

Z

ZZ

ZShaft Centerline

Pedestal Centerline

FIGURE 1. COORDINATE SYSTEM FOR THE API FORCES AND MOMENTS

Vertical In-Line Pumps Horizontal Pumps with Side Suctionand Side Discharge Nozzles

Page 92: Centrifugal Pumps Handbook

The Pump Handbook Series 93

charge pressure productto by-pass the dischargecheck valve and followthe discharge nozzle,then cross the top of thepump and exit at thesuction nozzle. Thismethod is inadequatebecause it heats onlythe top of the pump,resulting in a 150-300°Fdifferential within thepump. This tempera-ture differential createsa very large humpingtendency since the pump expandsmuch more at the top than the bot-tom. As a result, the bearing bracketsmight shift down at both ends of thepump. The rotor bows in the samedirection. In fact, the rotor will actu-ally roll over by hand about 70° untilthe wear rings rub. A few minuteslater, the heat in the top of the rotorwill allow it to roll about 70° further.

To assure that hot product is dis-pensed to the bottom of the pump,the product should be evenly distrib-uted in all suction and discharge cav-ities at the drain connections. Evendistribution provides the best oppor-tunity for thorough heat delivery tothe pump case, rotor, anddischarge/suction piping, prior topushing the start button. Slow rollingthe pump will also aid warm-up. Es-tablish the slow roll at 50-100 rpmand then initiate the warm-up flow tothe pump with hot product. • Piping supports

Bad installation, deterioratinghangers and foundation settlementare some of the most common causesof piping strain. Piping should bewell supported by spring hangers,anchors, expansion loops or compres-sion spring cans. In addition, all pip-ing supports should have adjustmentcapability to enable repositioning inresponse to deterioration and settle-ment. Glidden envisions a monitoringsystem, which would examine align-ment once a year or so, to test forchanges due to deterioration overextended periods.

The hangers and other supports,including the beams that supportthese hangers, should be designed tosupport the weight of the piping andthe process-liquid. If the support sys-

tem deflects, the pump may becomethe anchor for the piping.

Placing a firm anchor on a kneebrace right at the pump base mayprovide a good solution. The kneebrace, which should also be support-ed by spring hangers, forces the pip-ing up and away from the pump, sothat when the load changes as theprocess liquid flows into the piping, itwill relax down to the pump, and thespring hangers will bear the full load,allowing the pump to operate withvery low stresses.

CONCLUSION

The response to the discussiongroup at the Pump Users Symposiumindicates a real need for these kindsof practical solutions to nozzle load-ing. The pump operators present atthe Symposium recognized that aninexpensive piping design can becostly down the road. And, many arehopeful that the 8th edition of theAPI 610 standard, currently inprogress, will advance one step fur-ther toward an agreeable solution formore uniform nozzle loading prac-tices. ■

REFERENCES

1. Steiger, James E. ”API 610Baseplate and Nozzle LoadingCriteria,” Proceedings of the ThirdInternational Pump Symposium,(1985). pp. 113-129.

2. American Petroleum Institute,”Centrifugal Pumps for GeneralRefinery Services,” API Standard610, Seventh Edition (1989).

Kimberly Fortier is Assistant Editorfor Pumps and Systems.

CONFRONTING THE ROOT CAUSE

Understanding how the baseplate and piping design relate pro-vides one key to maintaining shaftalignment and thereby pump reliabil-ity. However, as Steiger maintains,”The pump-baseplate assembly rep-resents a complex structure whoseresponse to piping loads is difficultto predict with a high degree of cer-tainty.” Joseph agrees, ”While somepump cases and base plate founda-tion designs can take very highloads, there are others for whichsimply tightening a nut one flat at atime changes alignment significant-ly.” If the pump case and base plateconstruction is reasonably rigid,higher forces may be applied withlittle deflection at the coupling.”However,” Joseph warns, ”if usersdepend on the pump case and thebase plate to provide the rigidity,and the piping is significantly high instress application, then they’re justcovering one problem with anothersolution. They’re not getting at theroot cause—the piping strain.”

Joseph recommends properwarm-up procedures and piping sup-ports, in addition to good pipingdesign, as the primary remedy forpiping strain. His recommendationsare outlined as follows:

• Warm-up proceduresLarge, hot (and, as some users

have pointed out, expensive) pumpsare especially affected by nozzle load-ing, and for these pumps properwarm-up is essential. API recom-mends that the warm-up procedurebe very well thought out. Withoutproper warm-up the pump may suf-fer from uneven thermal growth inthe piping, case, shaft, stuffing boxes,and bearing housings. As the pumpcomes to equilibrium, it will experi-ence transient thermal growth whichmay put it under considerable stress.The design of the piping is also cru-cial to adequate warm-up andshould allow hot product to flowfrom the discharge line to the bot-tom of the pump case.

Many operators bring hot prod-uct to the pump by employing a by-pass to the discharge check valve.This is a small piece of pipe with ablock valve which enables the dis-

Shaft

Casing

FIGURE 2. RESULTS OF NOZZLE LOADING,COMPOUNDED BY HOT OPERATION

Page 93: Centrifugal Pumps Handbook

94 The Pump Handbook Series

rocess requirements oftendemand capacities belowthose achievable with a con-ventional centrifugal pump.

Figure 1 illustrates the range of ser-vice conditions considered to be lowflow. The minimum continuous sta-ble flow of a typical 1”x2”x7” over-hung pump at 1800 rpm is approx-imately 7 gpm, while at 3550 rpmthe minimum continuous stableflow is about 13 gpm. A pump ofthis size will produce about 240 ft.of head. As the head requirementincreases to 5000 ft., the minimumcontinuous stable flow will increaseto about 190 gpm.

Conventional centrifugal pumpswill not handle these low capacitiesvery well for two main reasons:

Suction recirculationThe minimum continuous sta-

ble flow is usually set by the pumpmanufacturer to avoid suctionrecirculation. Suction recirculationresults in increased vibration andimparts continuous axial move-ment to the shaft, decreasing thelife of bearings and mechanicalseals. The point at which suctionrecirculation begins may be calcu-lated as described by Dr. S.Gopalakrishnan in his presentationat the 5th International PumpUsers Symposium in 1988. Thepump manufacturer should per-form these calculations and set thepump minimum continuous stableflow at a capacity greater than thecalculated capacity.

Temperature riseThe ultimate limitation on low

capacity is minimum continuous

Low Flow OptionsThis service range demands an innovative approach.

thermal flow. Temperature risethrough a pump determines the min-imum flow rate. The maximum safetemperature rise through a pumpshould be limited to 10°F. The for-mula for determining thermal risethrough a pump is:

δT = H x 1 778Cp (Eff - 1)

H = total head in feet

Cp = specific heat of the liquid in Btu x °Flb

788 ft-lbs = the energy to raise the temperature of one pound of water by 1°F

PARTIAL EMISSION PUMPSThe type of pump most fre-

quently applied to fulfill low flow

requirements is a single port dif-fuser pump with a ”Barske” straightvane impeller close coupled to anelectric motor, also known as a par-tial emission pump (Figure 2).Theoretically, in this kind of pump,the only liquid discharged as eachchamber passes the diffuser port isthe liquid between the impellervanes. In reality, however, due tothe clearance between the case andimpeller, some additional liquid alsogets swept out the diffuser port.Unfortunately, this pump has a headcapacity that droops at shutoff whichinhibits the ability to control thepump capacity by increasing pres-sure with decreasing flow (Figure 3).As a result, installation of a flowme-ter is necessary to effectively controlthis type of pump.

BY PUMPS AND SYSTEMS STAFF

P FIGURE 1

CENTRIFUGAL PUMPS

HANDBOOK

Page 94: Centrifugal Pumps Handbook

The Pump Handbook Series 95

Because the characteristic curvefor the ”Barske” impeller, alsoreferred to as a high solidity impeller,is exactly the opposite of a centrifugalpump (where increasing the numberof impeller vanes will flatten thecurve and eventually cause a droop),the droop of the head capacity curvetowards shutoff can be minimized ina single port diffuser pump byincreasing the number of impellervanes.

Another method of eliminatinghead capacity droop is to install a dis-charge orifice. Since the frictionacross an orifice increases as the flowincreases, the pressure of a dischargeorifice will increase the pump curveslope so that the pump can be pres-sure controlled. Unfortunately, a dis-charge orifice decreases the pumpefficiency.

The partial emission pump isalso available with an integral gear(either single or double increaser) toproduce a higher pressure head thana single stage pump. Since high headapplication with the integral gearmay call for speeds up to 20,000 rpm,an axial flow inducer is oftenemployed in conjunction with thisgear to lower the net positive suctionhead (NPSH).

Another means to achieve lowflow combined with high headrequirements is to drive the pumpwith a special motor capable of highspeed. Application of a variable fre-quency drive will produce speedsnearing 7200 rpm. With this type ofconstruction a partial emissionpump may also be coupled to acanned motor for sealless pump con-struction.

One manufacturer builds thepartial emission type pump with anin-line configuration, giving thepump its own bearing frame. Inthis configuration the pump is flexi-bly coupled to a standard verticalsolid shaft motor. This same man-ufacturer also builds this pump in ahorizontal centerline-mounted con-figuration.

FLOW RESTRICTION DEVICESConventional centrifugal pumps

can handle low flow conditions withthe incorporation of a restriction

Pump Casing

Impeller

Conical Diffuser

Diffuser Throat

FIGURE 2. BARSKE STRAIGHT VANE IMPELLER WITH SINGLE PORT DIFFUSER PARTIAL EMISSION PUMP

“Barske” Impeller-WithSingle Port Diffuser

4 Vane ImpellerVolute Case

4 Vane / Volute

Efficiency % Curves

0

10

20

30

40

50

60

70

80

90

0

20

40

60

80

120

100

Gallons Per Minute

Efficiency %Total Head

In Feet

FIGURE 3. TYPICAL CURVE SHAPES

40200 60 80 100 120 140

Barske ImpellerSingle Port Diffuser

Head-Capacity Curves

Page 95: Centrifugal Pumps Handbook

96 The Pump Handbook Series

tions the pump will produceheads up to about 1000 ft. A pri-mary advantage of the verticalpump is the ability to stackmany stages so that a lowcapacity impeller of fairly goodefficiency will produce a highhead. These pumps usuallyincorporate two, sometimesthree or four, impeller designsof various capacities. Mixingimpellers will result in a ratedpoint capacity very near BEP.There is a limit, of course, tohow many stages a vertical canpump may have. The limitingfactors are shaft diameter sizerequired to transmit the horse-power and torque and the avail-ability of shafting in longsections (usually 20 feet).Another limitation is dependenton the machining tolerances ofthe register fittings of the bowlassembly. Since the tolerancesare additive as the bowl isassembled, they may causeshaft binding if they are nottight enough.

• Regenerative turbine pumpswill also fulfill low flow require-ments. These pumps, availablein single and multistage con-struction, have a very steephead characteristic and willoperate on pressure control. Theregenerative turbine does notdemonstrate any apparent prob-lems with minimum continuousstable flow, so the only limitingfactor to set minimum flow istemperature rise. The formulafor temperature rise throughthese pumps is identical to thatfor centrifugal pumps. The dis-advantage of a regenerativeturbine is the close internalclearances required to producethe pumping action. To accom-modate this close clearance, thepumpage must be very clean.

• Gear or other rotary positive dis-placement pumps also will oper-ate in the low flow range withoutdifficulty. These pumps do not,however, operate well in low vis-cosity services.

device on the discharge to shift thebest efficiency point (BEP) capacityback toward shutoff and increasethe pump curve slope. Unlike thepartial emission pumps whichemploy a construction requiringremoval of a motor with a specialshaft extension for mounting theimpeller, or in the case of highspeed applications removal of themotor and gear, the application offlow restriction devices on conven-tional API or ANSI pumps providesthe benefit of an easily maintainedsingle stage pump.

Even though a restriction devicereduces the efficiency by a consider-able amount, low flow pumps aregenerally low horsepower machines,so consuming a little more horsepow-er to obtain a steep curve rise previ-ous to shutoff is a small price to payfor the more desirable performance.Moreover, the required motor horse-power for the restricted pump is lessthan that for a non-restricted pump,as the restriction will not allow thepump to run to the extended portionof the curve.

The use of an orifice to restrictflow will produce the desired per-formance. However, if the orificediameter is considerably smallerthan the pump discharge and dis-charge piping, cavitation and noisemay occur on the downstream sideof the orifice.

For this reason one pump man-ufacturer incorporates a venturi tomodify pump performance. Theadvantage of the venturi is that thegradual taper down to the requiredhole size then back up to the dis-charge pipe size effectively elimi-nates the cavitation, noise andvibration. Pumps equipped withventuri have been observed to runsmoother and quieter as theyapproach shutoff.

OTHER PUMP OPTIONS• Another type of centrifugal

pump that will operate effec-tively in the low flow range is avertical can pump. A 5-6 in.diameter bowl assembly willexperience its BEP capacity inthe 60-120 gpm range at 3600rpm. At these operating condi-

• Controlled volume meteringpumps can be applied for lowflow services and are one of thefew types of pumps that willoperate at flow rates below 1 gpm. The disadvantages ofusing a metering pump are theinherent pulsations which maydamage downstream pipingand instruments. Pulsationdampeners help to smooth outpulsations but never entirelyeliminate them.

CASTING LIMITATIONSDevelopment of a truly efficient

low capacity centrifugal pumprequires prohibitively small liquidpassages. These small passages aretroublesome to produce in the cast-ing process because the sand mold isprone to collapse at such small sizesand small interior passages are diffi-cult to clean to the degree requiredfor good efficiency in operation.

A semi-open impeller is easierto cast and clean. This design is,however, in violation of API 610,which calls for an enclosed impellercast in one piece. If sufficientadvantages of the semi-open config-uration are demonstrated, this stan-dard might be changed. Very smallimpellers might even be machinedfrom billet stock (similar to somecentrifugal compressor impellers),thus eliminating all of the castingproblems.

Similarly, the casing of a lowflow pump is difficult to cast andclean, requiring very small passage-ways which must have a smoothsurface in order to produce goodefficiency. This obstacle to produc-ing a low flow pump case might beovercome by eliminating the needfor a case casting in favor of amachined and fabricated construc-tion.

EVALUATING HYDRAULIC FITThe fact of the matter is most

manufacturers usually make littleprofit on their small model pumps.To convince manufacturers thatquality low flow pumps are actuallyin demand, users must let themknow that their quotations forpumps in the low flow area are being

Page 96: Centrifugal Pumps Handbook

The Pump Handbook Series 97

evaluated for hydraulic fit. Onemethod of evaluating hydraulic fit isshown in Figure 4. This evaluatingtool adds a penalty, as a percentagemultiplier, to the pump price forrated capacity to the left of BEPcapacity. This tool is based on thefact that a higher suction specificspeed correlates with a smaller stablewindow of operation.

Applying this tool consistentlyand sending it along with yourrequest for quotations will convincepump manufacturers that low flowperformance is an area of hydraulicdesign that needs to be addressed. ■

REFERENCESF. H. Fraser, Recirculation in

Centrifugal Pumps, presented at theASME Winter Annual Meeting(1981).

S. Gopalakrishnan, A NewMethod for Computing Minimum Flow,presented at 5th International PumpUsers Symposium (1988).

U. M. Barske, Design of OpenImpeller Centrifugal Pumps, RoyalAircraft Establishment, Farnborough,Technical Note No. RPD 77 (January,1953).

Trygve Dahl, Centrifugal PumpHydraulics for Low Specific SpeedApplications, presented at 6thInternational Pump Users Sym-posium (1989).

ACKNOWLEDGMENTS Pumps and Systems would like to

thank the members of its UserAdvisory Team for their assistance inpreparing this article.

FIGURE 4

Page 97: Centrifugal Pumps Handbook

98 The Pump Handbook Series

roper lubrication is a key tolong, trouble-free life of cen-trifugal pump bearings. Inrecent years the issue of lubri-

cation has received renewed atten-tion from pump users in chemicalplants, pulp and paper mills, refiner-ies, and other industries.

Budgetary pressures have forcedmany plants to reduce maintenancecapital. Many knowledgeable mainte-nance workers have been laid off.Not surprisingly, the ability to main-tain pumping equipment properly isreduced, resulting in increased out-ages, lost production, and risingmaintenance costs.

Users have started to look topump manufacturers to pick up theslack and help solve pump reliabilityproblems, extend component life, andincrease mean time between betweenfailure (MTBF) and mean timebetween scheduled maintenance(MTBSM).

Statistics show (Ref. 1) that mostpump failures are related to bearingsand seals. In this article we will lookat bearings, analyzing how designchanges affect bearing life in a quan-tifiable way.

The need for improved pumpreliability and increased MTBF led toa new design, introduced by Gouldsin 1990/1991. Figure 1 shows crosssections of two single-stage, end-suc-tion ANSI pumps. Both have identicalwet ends (impeller and casing), butthe power end and the seal chambersare different. Improvements in the

seal chamber are signifi-cant. The new designhas a larger chamber toensure better heat trans-fer and cooler operationof the mechanical seals.The previous designincorporated a tightstuffing box.

The new power enddesign in Figure 1 fea-tures approximatelythree times the volumeof the oil sump (I), an oillevel sight glass (II) toassure the proper oillevel versus the constantlevel oiler (III), improvedcooling via a finned cool-er insert (IV) versus bot-tom cooling pockets (V),labyrinth oilframe seals(VI) versus lip seals (VII),and stiffer footing (VIII)for reduced vibrations.

A testing programhas been conducted tocompare the two de-signs under extremelyadverse operating condi-tions, such as runningendurance testing atoverspeed and belowminimum flow. Thisprogram was conductedat the R&D lab of theTechnology Center at Goulds, result-ing in quantifiable correlationsbetween changes in pump designand their effect on life extension.

Feedback from users comparing twodesigns was also obtained , specifical-ly in relation to the operating temper-ature of the bearing frame surfaces.

Pump Design ChangesImprove Lubrication

Quantifying the benefits of modifications.

Cross sectional views of old and new powerend designs.

BY LEV NELIK

P FIGURE 1

CENTRIFUGAL PUMPS

HANDBOOK

Page 98: Centrifugal Pumps Handbook

The Pump Handbook Series 99

approximately 13%longer life.INCREASED OIL SUMP DEPTH

A deeper sumpallows contaminantsto settle farther frommoving parts, result-ing in a cleaner layerof oil near the ballbearings (Figure 2).Contamination of thebearing races and theballs is the cause ofmicroscopic deteriora-tion of load surfaces,

leading to failure. Statistics show (Ref.2, 4) that a cleaner oil operation canincrease bearing life by nearly 2.1times (Ref. 3). Similarly, due todecreased air concentration, the oil

oxidation rate by air is reduced forthe larger sump. Again, for the typeof pump studied in this work, thisresults in a 2% extension in bearinglife (Ref. 3).LABYRINTH OIL VERSUS LIP SEALS

The effects of oil contaminationare further reduced by improved oilseals. A proprietary labyrinth sealdesign was tested against the lip seal.Both pumps were sprayed with waterfrom a hose, simulating plant wash-down. The spray was directed at vari-ous angles to the frame at the oil andseals area. The oil was then analyzedfor water content. It was found thatthe previous design equipped with lipseals contained 3% water after 30minutes of spraying, while the newdesign, with labyrinth seals, showedno water at all. Also, lip seals may

ANALYSIS OF THE POWER ENDWith regard to the power end

(Figure 1), the belief that “the biggerthe better” is not uncommon in thepumping community. This idea hassome merit, but manufacturers oftenoverlook the importance of quantifyingthe benefits of a particular design ormodification. Frequently, little infor-mation is given as to how much lifeextension can be obtained by, say,having a deeper sump, or how muchadded value and savings can be real-ized from the increased bearing frameheat transfer surface.

It is clear that a systematicapproach to identify, measure, andimprove pump component design isimpossible without a proper balanceof theory, experimentation, userfeedback, and data from real worldinstallations. Theory and experimen-tation should be balanced by clearcommunication between manufac-turers and users.INCREASED FRAME OUTSIDE HEATTRANSFER SURFACE

Heat is transferred from thepump bearings to the oil and throughthe housing frame walls to the outsideair. Some of the heat is also conduct-ed through the casing to and from thepumpage, depending on the tempera-ture of each. Typically, the differencein temperatures is small for thepumping conditions of chemicalplants, and the effects are omitted forsimplicity.

Our investigation has shown(Ref. 3) that the larger surface areacan result in a nearly 40°F reductionin bearing operating temperature. Thecooler bearings, in this case, result in

FIGURE 2

Larger sump results in reduced concentrationof contaminants, which settle to the bottom.

FIGURE 3

Comparison between bearing submergence in oil, operating temper-ature, and bearing life for the old and new designs.

Page 99: Centrifugal Pumps Handbook

100 The Pump Handbook Series

cause wear and leakage after approxi-mately only 2,000 operating hours.

OIL LEVEL SIGHT GLASS VERSUSCONSTANT LEVEL OILER

A large sight glass allows directvisual observation to ensure proper oillevel. It is standard in the new design,although a constant level oiler optionis available. A constant level oiler ispreferred by many users. When prop-erly installed and maintained it canresult in satisfactory operation.

However, because operation ofthe oiler is “blind,” depending solelyon strict conformance to correct (andnontrivial) oil filling and maintenanceprocedures, it may lead to an incorrectoil level inside the frame. This canlead to hot operation and prematurefailure. Another problem is known asthe oiler “burping” effect, resulting in ahigher actual oil level than perceived(Ref. 5). Obviously, the new pumpdesign can be equipped with both thesight glass and oiler if they are desiredby the user.

Such improvements in designcan be combined because they bene-fit pump reliability independently.Based on this research, when all areadded together, an improvement inpump life of up to 125% may beobtained. Even longer life may berealized due to other design upgrades,such as providing the pump with amore rigid foot, reducing vibration,improving (finned) cooling, and creat-ing larger chambers for mechanicalseals. For brevity, these effects arenot included in this analysis, but theycan be accounted for in the refer-ences (Ref. 3, 5).TEST PROGRAM

To support and validate the theo-retical derivations and assumptions asoutlined below, a testing program wasconducted, including lab testing andfield data analysis. Figure 3 shows acomparison between operating tem-peratures and bearing life for the old

and new designs. Tests were conduct-ed with oil covering different levelsof the lower ball of the bearings. Theproper design level (marked 50% onFigure 3) corresponds to oil at themiddle of the lower ball of the bear-ing. At design setting, the new frameran 40°F cooler, with a correspond-ing predicted life extension ofapproximately 6,000 hours.ECONOMIC BENEFITS

Having measured the lifeincreases resulting from theseimprovements, it is not difficult toassess the economic benefits of thenew design.

Assuming the average value forMTBF of two (2) years for the olddesign, a 125% improvement resultsin a four and a half (4.5) year bearinglife for the new design. The recipro-cals of these numbers (1/2 years =0.5; 1/4.5 years = 0.22) give anapproximate number of failures orscheduled maintenance per year.The difference, 0.28, when multi-plied by the average cost of repair of,say, $260 in parts and labor and3,000 pumps per plant results in ayearly plant savings of:

0.28 x $260 x 3,000 = $218,400In addition, savings resulting

from increased uptime and a reduc-tion of lost production at approxi-mately $500 per off-line hour,assuming an average four hours perrepair for off-line time, would be:

0.28 x ($500 x 4) x 3,000 = $1,680,000The total, $1.9 million, is annual

plant maintenance savings.Obviously, these numbers are approx-imate and can best be determined byindividual maintenance departmentsusing their operating specifics, but thesavings potential due to improveddesign is clear.CONCLUSIONS AND RECOMMENDATIONS

Our study demonstrated thatsubstantial savings can be realized

through improvements in pumpdesign. To gauge such improvementssystematically, it is imperative toquantify the benefits of each pumpenhancement.

It is also important to maintain aproper balance between the solidtheoretical foundations used for theanalysis and the laboratory work,field testing, and data supportingsuch theory. Users should seekquantitative data demonstratingimprovements from pump manufac-turers, including improvements inMTBF and MTBSM, enabling themto determine added value and othereconomic benefits.

This approach will improvecommunication between manufac-turers and users, and lay the groundwork for the next step: furtherimprovements in pump reliability.REFERENCES1. H. Bloch. PRIME I and II, Pump

Seminar Series, 1992/1993.2. SKF General Catalog 4000 US (bear-

ings), 1991.3. L. Nelik. Value Added and Life

Extension with Regard to Reliabilityof X-Series 3196 ANSI Pump. GouldsPumps, Inc. Internal Report, 1993.

4. CRC Handbook of Lubrication, Vol. 1,CRC Press, R. Booser, 1983.

5. L. Nelik. Goulds Technology VideoSeminar. Constant level oilers versussight glass, Series 0693-01.

For more information on these refer-ences please call (315) 568-2811. ■

Lev Nelik is Manager of PumpTechnology for Goulds Pumps. Hisresponsibilities include developmentalwork in various aspects of centrifugalpump technology, developing new prod-ucts, and improving the reliability ofexisting products. Dr. Nelik hasauthored publications on centrifugalpumps and hydraulic power recoveryturbines, fluid mechanics, heat transfer,and FEA CAD/CAM applications.

Page 100: Centrifugal Pumps Handbook

The Pump Handbook Series 101

oday chemical manufacturersand users are faced with globalcompetition and pending envi-ronmental restrictions that

threaten to reduce profitability. Theneed to reduce overall operating costshas driven pump users at chemicalplants to focus on improving reliabili-ty and eliminating or reducing fugi-tive emissions.

SEALED PUMPS

The mechanically sealed chemi-cal process pump, which meetsASME/ANSI B73.1M standards, is theworkhorse of chemical processingindustries. It will continue to be usedon a wide range of process applica-tions—such as liquids containing sig-nificant amounts of solids (sodiumchlorate, alum, sodium carbonate,chemical wastewater), light slurries(silver nitrate and acetone slurries),viscous liquids (above 150 cP, includ-ing black liquor and titanium diox-ide), and stringy materials wheresealless pumps may not be economi-cal to use. In addition to its ability tohandle tough services, the flexibilityof the design—along with improvedlow-emission mechanical seals—con-tinues to make ANSI pumps the stan-dard in this field.

To elaborate on why seallesspumps are not economical to handlethe above materials, we must notethat they use enclosed impellers toreduce the axial thrust and increasereliability. (Although several manu-facturers have tried using open orsemi-open impellers in seallessdesigns, many of these have not beenreliable at two-pole speeds.) Also,standard sealless pumps have smallinternal passageways to circulateliquid for bearing lubrication anddrive-end cooling, and mechanicalseal manufacturers are rapidlyimproving the reliability of their prod-

ucts while reducing emissions to wellbelow 500 ppm. With this in mind,consider the following:1. Enclosed impellers are prone to

plugging and premature wear inthe above services due to smallwear surface area. (Performanceand efficiency cannot berenewed without replacing wearrings.)

2. Open or semi-open impellers arereliable in these services and arestandard for ANSI pumps. (Simpleexternal impeller adjustmentsallow easy maintenance of perfor-mance and efficiency, and thereare no wear rings to replace,yielding long-term energy sav-ings.)

3. The small internal passagewaysin sealless pumps are subject toplugging while handling liquidswith only small amounts (5%) ofsolids. Viscosity handling is alsolimited.

4. Design solutions separate thepump end from the drive end toallow sealless pumps to handlethese services, but these modifi-cations can be expensive andmay not be cost effective.

Considering all the facts, it’sunderstandable that mechanicallysealed ANSI pumps are the moreeconomical choice to handle thesetypes of liquids.

ANSI RELIABILITY IMPROVEMENTS

To meet emissions regulationsand improve reliability, processindustries have pushed ANSI pumpmanufacturers to improve perfor-mance. Some manufacturers haveformed alliances with users to sharetechnology and improve standarddesigns. By working together, the the-oretical has been combined with the

realities of applying pumps on a mul-titude of services and making themlast.

These efforts have produced thefeatures listed below that manymajor ANSI pump manufacturershave incorporated (Figure 1). At aminimum, users should purchaseANSI pumps with features that bestmeet their application needs.However, most new designs incorpo-rate features systematically to provide reliable products. Com-promising designs to save money oradd standard plant features—substi-tuting a vendor’s standard labyrinthseal with the plant’s standard oil seal,for example—may not be advisable.New ANSI pump features includethe following:

1. Labyrinth oil seals are designedto prevent premature bearingfailure from lubricant contami-nation or oil loss. These non-con-tacting seals have replacedBuna-rubber lip seals, whoseuseful life was three to sixmonths under normal condi-tions. Materials of constructioninclude carbon-filled Teflon,bronze, or stainless steel.

2. Increased oil sump capacity pro-vides better heat transfer formore effective oil cooling.Bearings operating at lower tem-peratures contribute to longerlife.

3. A rigid frame foot reduces theeffect of pipe loads on shaftalignment. Misalignment won’texceed 0.002 in. under load, andpump and driver alignment isbetter maintained.

4. Bull’s eye sight glasses insureproper oil level, which is criticalto bearing life. Level oilers haveoften been misused, leading to

BY RICHARD BLONG AND BOB MANION

T

Increase reliability and reduce emissions through pump selection.

CPI Pumping

CENTRIFUGAL PUMPS

HANDBOOK

Page 101: Centrifugal Pumps Handbook

102 The Pump Handbook Series

over- or under-filling sumps,both of which contribute to bear-ing failure. Sight glasses are alsoconvenient for checking the oilcondition visually to determine ifa change is necessary. Constant-level oiler manufacturers are justnow introducing oilers that elimi-nate the potential for improperoil level settings while providinga sight glass, combining the bestfeatures of both methods.

5. Mounting flanges accommodatean optional adapter that simpli-fies pump/motor shaft alignment,saving the user time and moneyduring installation.

6. Condition monitoring bosses onpower ends provide consistentmeasurement points for tem-perature and vibration sensors.Many users report increasedpump life from using predictivemaintenance to identify andcorrect problems early. Takingmeasurements at the samepoint aids in proper interpreta-tion of readings and allows per-sonnel to move through theplant more quickly on inspec-tions.

7. Engineered large seal chambers,specifically designed for today’smechanical seals, increase seallife through improved lubrica-tion, cooling, air venting, andsolids handling. The chambersallow seal manufacturers toengineer and apply more reli-able designs, including cartridgeseals.

These developments extendpump and seal life and reduce emis-sions at the same time. Experienceshows that one cannot be accom-plished without the other. For exam-ple, a mechanical seal with emissionsin excess of regulations has alreadyfailed in its application.

Another benefit of these featuresis that several manufacturers and sealsuppliers are extending unconditionalwarranties to as long as three years,helping to further lower operatingcosts.

SPECIALTY PUMPS FOR IMPROVEDRELIABILITY

Many diversified chemical pro-ducers are moving production of com-modity chemicals to the Asia-Pacificand Latin American regions to take

advantage of lower labor and produc-tion costs. As a result, chemical pro-duction in the United States is beingdriven toward manufacturing special-ty chemicals typically produced insmall runs or batches. Examplesinclude methylisobutyl- ketone(MIBK) and paratertiarybutyl- phenol(PTBP). Pump applications for thesebatch-type processes are usually lowflow, in the range of 0 to 100 gpm.

Traditionally, users install stan-dard process pumps and throttle thedischarge valves to obtain low-flowperformance. However, these pumpsare not designed to operate continu-ously in this range (Figure 2). Higherradial loads and increased shaftdeflection lead to premature bearingand seal failure. Costly downtimeand maintenance expenses result.

For low-flow operation, usersshould specify a pump designed tomeet specific service conditions(Figure 3). ANSI pumps designed forlow-flow operation are available toincrease pump and plant reliability.

Improvements come from acasing and impeller designed forlow-flow operation. Low-flowdesigns use concentric volutes andradial vane impellers to reduce radi-al loads, eliminating hydraulic andmechanical problems from throttledlow flows (Figure 4). Some designsreduce radial loads as much as 85%compared to end-suction expandingvolute pumps in this service (Figure5). Shaft deflections from high radi-al loads are minimized, optimizingbearing, mechanical seal, and over-all pump life. A disadvantage oflow-flow ANSI pumps is that theysacrifice some efficiency to reliablyhandle viscous and solids-contain-ing liquids.

Another approach to lowflow–high head applications is theregenerative turbine pump. Thisdesign directs liquid by a passagewayso that it circulates in and out of theimpeller many times on its way frompump inlet to outlet. Both centrifugaland shearing action work together toefficiently develop relatively highheads at low flows. Regenerative tur-bines also use concentric volutes andradial vaned impellers to obtain thereliability benefits discussed above.One drawback is that this type ofpump utilizes close running clear-

ANSI pump improvements.

FIGURE 1

Page 102: Centrifugal Pumps Handbook

The Pump Handbook Series 103

ances to keep efficiency high and it istherefore normally used on clean liq-uid applications.

SEALLESS PUMPS

With the implementation of theClean Air Act, sealless pumps offer adynamic solution to controlling emis-sions. Not only should sealless pumpsbe strongly considered to controlemissions of the 149 volatile organiccompounds identified by the En-vironmental Protection Agency, butthey should be viewed as solutions tomany difficult applications encoun-tered in CPI plants today. For exam-ple, if users are experiencing sealingproblems because of the pumpedproduct’s poor lubricity (typical ofacidic products in the range of 0–3pH, such as sulfuric or hydrochloricacids), difficulty with product crystal-lization at seal faces (usually withcaustic products in the range of 10–14pH, such as sodium hydroxide andpotassium hydroxide) or are frustratedwith sophisticated auxiliary pipingplans to provide clean, cool flush liq-uid to mechanical seal faces, seallesspumps may be the answer.

IMPROVED RELIABILITY WITH MAGDRIVES

It is well recognized thatmechanical seals and bearings are the

two items that failmost often inpumps. These fail-ures are oftendirectly related toimproper applica-tion and installa-tion, pooroperating prac-tices or lack ofmaintenance, pipestrain, or misalign-ment. All of theseagain lead to highbearing loads,shaft deflection,and bearing andseal failure.Magnetic drivepumps have nei-ther a mechanicalseal that can failnor a driven shaftthat can be sub-jected to pipe

strain or misalignment. The drivenshaft is separated from the drive shaftby a magnetic coupling, eliminatingthe two major causes of pump failure.

CRITICAL MAG DRIVE FEATURES

Reliable magnetic drive pumpsmust address two critical concerns:

1. proper lubrication of the journalbearings

2. removal of heat generated byeddy currents in the recirculationcircuit

The design must deliver liquid tolubricate the bearings—it should notbe flashing or have risen in tempera-ture, which decreases lubricity, pre-vents proper cooling, and leads tobearing failure. Proper journal bear-ing lubrication directs cooling liquidto the bearings, then to the magnets.Dual path designs provide lubricationto these areas separately. Bothapproaches prevent flashing at thebearings, a leading cause of failure.

Another typical mag drive pumpfailure is liquid flashing at theimpeller eye after being circulatedthrough the drive end to removeeddy current heat. The result is avapor-bound pump. New mag drivedesigns have virtually eliminated thisproblem by creating a constant pres-surized circulation circuit that pre-vents flashing of cooling liquid andthe associated failures (Figure 6). Notall new designs use pressurized circu-lation, and because most regulatedliquids are volatile, this feature is nec-essary to achieve extended life inthese services.

TO

TA

L D

YN

AM

IC H

EA

D

TYPICAL END SUCTION PUMP CURVEFLOW

System Curve -ActualThrottled Operation

Rated Performance

FIGURE 2

Off-design (throttled) operation range (darker gray)and recommended operation range (gray).

Pump curve for a low-flow ANSI pump.

TO

TA

L D

YN

AM

IC H

EA

D

FLOW

System Curve -ActualThrottled Operation

FIGURE 3

Page 103: Centrifugal Pumps Handbook

104 The Pump Handbook Series

Regardless of the design featuresand modifications available frommanufacturers, users are responsiblefor providing suppliers with as muchdata as possible on fluid and operat-ing conditions. To apply seallesspumps properly, many factors mustbe considered:• Is the flow continuous or inter-

mittent?

• Upon shut-down, what reaction(if any) will the process fluidhave to residual heat? Chemicalslike butadiene and formaldehydemay polymerize, leaving depositsinside the drive section and onthe bearings.

• Can the process shut down auto-matically, resulting in the pumpoperating at shut-off condition?

• Conversely, can the system allowthe pump to operate at theextreme right of the pump curve,which can adversely affectNPSHR and cause motor over-load or excessive thrust?

• What are the fluid characteris-tics, including vapor pressurecurves, specific heat, viscosityover the process temperaturerange, and the effects of heatingand cooling on the process fluid?Benzene freezes at 42°F (depend-ing on the installation location,address the possibility of expo-sure to low temperatures), andtoluene diisocyanate freezes at72°F and begins to polymerize at127°F (again, protect the installa-tion or use jacketing if neces-sary). Maleic anhydride freezesat 130°F (use heating jackets ortemperature control).

• What about the customer’s practi-cal knowledge of the corrosivenature of the chemical?Sometimes the standard corrosioncharts don’t give the whole story.

ABRASIVES

When pumping fluids containingparticles, the traditional solution is touse very hard bearings (silicon car-bide) operating against a hard or coat-ed journal. The application of seallesspumps should go beyond this seem-

ingly easy approach. Considerationmust be given to:• the abrasiveness of the solids

• the size of the particles

• the quantity of particles

• whether they can agglomerate

• what creates the particles (reac-tion, catalyst, temperature)

The size of the particle that canbe handled is usually determined bythe impeller design and the clear-ances in the fluid passages. Theeffects of the quantity of particles areusually predicted from previous expe-

rience and test data. Solids may beformed by reactions to moisture (tita-nium tetrachloride), temperature(butadiene or formaldehyde), or a cat-alyst (any process that uses a catalystthat may vary in quantity or is sub-ject to upsets).

When the fluid is understood, itmay be best to use modifications,including: backflushing to keep parti-cles out of the drive section, heatingor cooling jackets, heat exchangers inflush lines, filters or speciallydesigned units that utilize isolationchambers, built-in seals, and preci-sion back-flushing to reduce processstream dilution, if economical.Otherwise a mechanically sealedANSI pump may be the best solution.

MAG DRIVE CONDITION MONITORING

Magnetic drive pump reliabilityis also affected by operating practices.Condition monitoring devices can beapplied to shut pumps down before acritical failure. Maintenance can thenbe performed, or operator errors cor-rected, before the pump is put backinto service.

Temperature detection and powermonitoring together provide the bestbasic protection. Temperature detec-tion indicates internal pump problemssuch as plugged recirculation paths,while power monitoring prevents dry-run failure. Other devices availableinclude low amp relays, leak detectionindicators, and package control sys-tems.

INSTALLATION

The effort involved in selectingthe right pump for a given CPI appli-cation can be nullified by poor instal-lation. As much effort, if not more,should be put into installation designto insure expected performance isachieved. (To understand how properprocedures improve equipment relia-bility see “Installation and Start-UpTroubleshooting,” Pumps and Systems,November 1993.) Important stepsinclude:

1. Lay out suction piping to provideNPSH available to the pump inexcess of NPSH required. A com-mon recommendation: NPSHA> NPSHR + 2–5 ft. See “PumpSuction Conditions,” Pumps and

FIGURE 4

An expanding volute pump(top) and a circular volutepump with a radial vaneimpeller (bottom).

Page 104: Centrifugal Pumps Handbook

The Pump Handbook Series 105

Systems, May 1993 and “HowMuch NPSH Is Enough?”September 1993.

2. Provide a straight run twice thelength of the pipe diameter (2D)to the pump suction flange toprevent added turbulence at theimpeller eye, which could lead topremature (incipient) cavitation.

3. Install conventional or cartridgemechanical seals according tomanufacturer recommendations.

4. Meet seal flush requirements byproviding an external flush at thenecessary pressure and tempera-ture, or add auxiliary piping forflushing on the pump.

5. Prepare the foundation beforegrouting the baseplate.

6. Select grout that will meet instal-lation requirements.

7. Select a baseplate to maximizepump, seal, and motor reliability.Many vendors offer baseplateswith enhancements such as .002in./ft flatness, leveling screws,motor alignment screws, continu-ous drip rims, and other featuresdesigned to ease installation andalignment and increase pump life.

8. Align equipment according tomanufacturer specifications.

9. Select and install condition moni-toring devices for sealless pumps.

CONCLUSION

Selecting a pump to improve relia-bility will reduce emissions and operat-ing costs at the same time. Neither amechanically sealed ANSI pump nor a sealless pump can be univers-ally applied on everyprocess application.Make an informed deci-sion based on specificservice conditions andtotal cost (initial + maintenance + operat-ing costs). To insure areturn on investment, asmuch time and effortmust be expended onthe design of equipmentinstallation as on pumpselection. Althoughselecting equipment forincreased reliability andreduced emissions mayseem expensive in theshort term, it savesmoney in the long run.Work with manufactur-

ers’ reps and rely on their expertise,but be informed, as well, and togetheryou can apply pumps properly in yourfacilities. ■

Rich Blong is product manager forchemical pump development for GouldsPumps Inc. Previously, he was a seniorapplications engineer responsible forapplying chemical pumps for many dif-ferent processes. He also worked as apump systems engineer with UnionCarbide’s Linde Division, now PRAX-AIR. He has a bachelor’s degree inchemical engineering from theUniversity of Buffalo.

Bob Manion is product managerfor magnetic drive and non-metallicpumps for Goulds Pumps Inc. He hasheld marketing and sales managementpositions related to developing, selling,applying, and servicing centrifugalpumps for 13 years. Mr. Manion holdsa bachelor’s degree in marketing fromthe Rochester Institute of Technology.

Expanding Volute

Circular Volute

85%Reduction

0 50 100 150 200Low Flow Operating Range–GPM

Incr

easi

ng R

adia

l Loa

d

FIGURE 5

Radial load curves.

FIGURE 6

Recirculation circuit.

Page 105: Centrifugal Pumps Handbook

106 The Pump Handbook Series

When the canned-motor pumpis the choice to solve a specificpumping problem and controlcosts, the following points must beconsidered to achieve satisfactoryresults:

CHEMICAL CHARACTERISTICS

There is seemingly neverenough information available onthe chemicals to be handled. Thesupplier must depend on the cus-tomer to provide this information,but it is also very important thatthe supplier and the customerexhaust their resources in anattempt to anticipate what a chemi-cal will do inside the seallesspump. Will it cause corrosion, boil,decompose, freeze, or polymerize?Any of these properties can resultin rapid failure unless anticipated.

APPLICATION AND METHOD OFOPERATION

Will the pump be used fortransfer, condensate return, reboil-er, or batch operation? Will it berunning continuously or intermit-tently? Will the location be remote,exposed to the elements, or in a haz-ardous location? How will thepump be operated and what willthe process demand? Can the flowrange over the complete curve? Is itclose to shut-off, which may requirea by-pass orifice? Or, conversely,will it occasionally pump at theextreme right of the curve? This canresult in cavitation and subsequentfailure if allowed to continue. All ofthese factors, combined with theknowledge of the fluid pumped,will determine the proper selectionand modifications necessary for suc-cessful pump operation.

DIAGNOSTICS AND CONTROL

Once the above factors havebeen determined, the user and sup-plier should agree on the type ofdiagnostic devices and process con-trol that will assure a successful installation. Diagnostics availableinclude:

• bearing wear monitors• rotation indicators• motor diagnostic devices• bearing temperature sensors• leak sensors• flow sensors

Any of the above may be rec-ommended. The pump and motorcan also be fitted with a controldevice such as:• a water or steam jacket• a water-cooled heat exchanger• a heat exchanger in the circu-

lation line• complete jacketing of the

pump and motor(Consider if the insulation will

create motor heat problems.)

MAINTENANCE

The final consideration must bemaintenance. Does the user have aplanned maintenance program?Does the user’s and supplier’s expe-rience indicate more frequent main-tenance intervals than normal withthe chemical product in this particu-lar mode of operation?

Proper maintenance andreplacement of less expensive bear-ings and gaskets can prevent amajor failure and yield increasedsavings.

CONCLUSIONUsing a sealless pump can be

easily justified due to the elimina-tion of leakage and emissionsbecause the value of the chemicallost using a sealed pump can becalculated. But there are manyother factors that are more difficultto quantify, including housekeep-ing costs, safety, odor, and publicand employee relations. The majorelements leading to long-term sav-ings using sealless pumps is the up-front analysis of the applicationand the supplier’s knowledge of hisproduct. ■

Joe Cleary is the Vice President(retired) of Sales for Crane Co.,Chempump Division.

Canned Motor Pumps

Page 106: Centrifugal Pumps Handbook

The Pump Handbook Series 107

BY: J.T. MCGUIRE

t first glance, pump require-ments don’t really seem thatcomplicated. After all, apump only needs to:

• move a specified volume of liquid through a given system

• be energy efficient• comply with any applicable laws

regarding leakage• achieve certain mean time be-

tween overhauls and replace-ment

• be delivered on time with complete documentation

• and all at minimal cost

It may sound simple at first, butit’s not. For example, some pumppurchasers may not know the vol-ume of liquid their system handles.Reliability data is hard to come by,too. With so many factors affectingpump life, mean time between over-hauls and replacement may not beknown. And the goals are conflicting.Increased service life may alsoincrease energy consumption andpurchase price. How do you sortthrough these factors? How do youdetermine what you need from apump, develop a meaningful specifi-cation for those requirements andfinally buy the right pump? Theanswer is to take it one step at a timeand follow a disciplined approach topump specification and purchasing.

DETERMINING PUMP REQUIREMENTSTo write requirements for a

pump, you should review the basicsof pumps. Pumps are designed tomove liquid against a hydraulic gradi-ent; in other words, to move liquidfrom the suction reservoir to the dis-charge reservoir, which differ in ele-vation and/or pressure (Fig. 1).

You can see immediately fromthe figure that the pump must supplyadequate energy to overcome the dif-ference in elevation and pressurealong with the friction losses in theconduits on both sides of the pump.It’s obvious that the pump, as thesole source of energy in the system,must supply all the needed energy.

A Thus it’s no wonder that severe con-sequences await those who overlookthat simple fact!

Another, perhaps not so obvious,fact is that the energy available at thesuction side must provide a certainnet margin over the liquid’s vaporpressure at the pump suction. Thisnet margin, called Net PositiveSuction Head Available (NPSHA), isnecessary to prevent cavitation — theboiling of liquid in the system.Cavitation impairs pump perfor-mance and shortens the service life ofthe pump. An excessive amount ofboiled-off vapor impairs themachine’s hydraulic performance. Inaddition, the subsequent collapse ofthe vapor bubbles as they move toregions of higher pressure can causecavitation erosion.

To prevent these problems, youmust specify total system head accu-rately. In most applications, you candetermine the normal pump flow andthe static components of the totalhead associated with ideal operationof the plant or process at its designoutput. Add estimated piping frictionlosses and control valve pressuredrop (if applicable) to find the totalsystem head for that capacity.(Remember friction head varies asthe square of the flow ratio.)

Normal pump flow and staticcomponents aren’t the whole story ofoperating conditions. You must alsofactor in the range of operating condi-tions your pump will be called on toperform under. Changes in operatingconditions can be caused by:• process unit downturn• flow swing to cover upset or

transient• change in static head as vessel

levels or pressures in bothchange with time

• change in friction head as systemfouls or scales or as dischargevessel fills

• pump wearYou use these data to compute

the rated flow — the flow underwhich your pump will need to oper-ate. You then match the performance

To get the right pump, all

you have to do is decide

what you want, state those

requirements clearly and

place your order with a

capable manufacturer.

Sounds easy, doesn’t it?

SYSTEMP1

P2

NPSH

PUMPINGENERGY

VAPORPRESSURE

HYDRAULIC GRADIENT

1

4

3

2

3

3

4

4

12

2

1

1

ENERGY LEVELS

- EXIT FROM SUCTION SOURCE

- PUMP SUCTION

- PUMP DISCHARGE

- DISCHARGE POINT

FIGURE 1

CENTRIFUGAL PUMPS

HANDBOOK

Pump Buying Strategies

Page 107: Centrifugal Pumps Handbook

108 The Pump Handbook Series

along with the total flow to behandled, the total head to bedeveloped, or the total powerabsorbed. It can also incorporatethe physical size of the pump(often related to type of pump)and the required turndown in flow when rated flow is high.

• Service lives of the pump andvarious components (Items 12through 15). These requirementsare usually expressed as meantime between failure. As notedearlier, data on the life of variouspump components are meager. Thus, these requirements are often not specified. Generally,antifriction bearings and the firststage impeller of high energycentrifugal pumps are the onlycomponents for which minimum service lives are commonlyspecified.

• Materials of construction (Item 16). You’ll have to handle this item since the pump manufac-turer does not control the pump-ed liquid. If you have little or noexperience pumping the partic-ular liquid, manufacturers willsuggest possible materials. Butthe only guarantee a pump manufacturer makes for materials is that they will conform to their producing specification.

• Extent of supply is an essentialissue. (Items 7, 8, 22, 23, 25 and26). When faced with increasingcomplexity and extent of specifi-cations, many pump purchasersfind it beneficial to summarizeextent of supply (also known asterminal points or battery limits)in a list or diagram. That’s a goodidea. It helps you state more clearly what you need.To simplify the specification, you

should note all technical elements ona basic data sheet. And rememberthat the basic data sheet should bejust that — a sheet. Multi-page datasheets are unwieldy. If your system iscomplicated, cite and add supplemen-tary sheets rather than cluttering thebasic data sheet.

Instead of building custom specs,some purchasers in particular indus-tries use general specifications issuedby that industry (for example, ANSIB73.1M-1991, which addresses hori-

Formats for the specificationsrange from a very simple functionalspec to a very elaborate functionaldesign and manufacture. The simplefunctional spec states only what thepump will be called on to do. Yougive complete freedom to the manu-facturer for designing the pump. Anadvantage to such a specification isthat you can get very interestingdesigns for unique pumping prob-

lems. But, such specscan be difficult to writeand you’ll need to evalu-ate the engineeringbehind the bids careful-ly.

To avoid the back-end expenses of a simplefunctional specification(and since most pump-ing requirements are rel-atively straightforward),most pump buyers writedetailed functionalrequirements and manu-facture specifications.

As a start, thesespecifications must address:• operating environment• liquid to be pumped• pump performance and life• materials of construction• extent of supply

Many other items related tofunction, design and manufacture canbe addressed (Table 1). The numberof requirements you choose toinclude will depend on the pumpingapplication and your confidence inthe potential pump manufacturers.

Items you may want to pay spe-cial attention to include:• Degree of redundancy (Item 4).

This item refers to the proportionof spare capacity the pumpingarrangement has to provide in the event one pump is lost. Typical values are 100, 50, 25 or0 percent. Most purchasers havedesign standards relating degree of redundancy to the type of service involved.

• Type of pump (Item 7). This issue is complex, determined by the hydraulic duty, the degree offlow regulation required, and the nature of the pumped liquid.

• Number of pumps (Item 7). Thisitem incorporates the requireddegree of redundancy (Item 4)

data (power, NPSHA, speed) quotedby the manufacturer against the ratedflow.

You can set the rated flow to themaximum rate at which your pumpwill be called on to operate. But if theflow range is very wide (and you planto use a centrifugal pump), you mightset the rated flow to the most fre-quent or efficient flow rate. In eithercase, once you set the rated flow andrequired operating flowrange, you will need tolook at the NPSHA forthe pump at these flows.

In addition to therated flow, you mustconsider the range offlow. Pumps cannotoperate across the entirerange of flow from max-imum flow to zero flow.With the exception ofdirect acting steampumps, no pump has aninfinite range. Centri-fugal pumps, the mostcommon pump in usetoday, can operate under a widerange of flows if they are designedappropriately. Thus, set the flow mar-gin to allow for process transients andpump wear, but don’t set it largerthan necessary. And be sure the stat-ed NPSHA reflects a value normallyavailable, not some possible mini-mum value with a hefty hidden mar-gin. Otherwise, you’ll find yourselfwith an unsuitable pump — an over-sized pump or one designed forabnormally low NPSHA.

DEVELOPING A MEANINGFUL SPECIFI-CATION

Once you determine the specifi-cations for your pump, you need tocommunicate those specs to the man-ufacturer in concise terms.

For the manufacturer to under-stand what you really want, yourspecification must state, in an orderlymanner, all the requirements youhave for the pump. But that doesn’tmean you should strive for a thickspecification document. The value ofthe specification is not proportional tovolume or weight. If anything, theinverse is true. Overly long specifica-tion documents often fail to statewhat the purchaser really wants.

Stated NPSHAshould reflect avalue normallyavailable, notsome possibleminimum value

with a heftyhidden margin.

Page 108: Centrifugal Pumps Handbook

The Pump Handbook Series 109

The old catchall, “comply withapplicable local, state, and federalrules and regulations,” doesn’t addanything to the specification.

Beyond technical requirements,the specification also must addressthe proposed terms of purchase orcommercial terms and conditions.Although these items are generallythe province of the purchasingdepartment, you, as a specifying engi-neer, should be aware of what isinvolved. The major items covered inthe terms and conditions are:• delivery period or date• point of delivery• liquidated damages• terms of payment• warranty• default

• cancellation• bankruptcy

Delivery period and terms ofpayment should be of special interestto you. For example, if the deliveryperiod is too short, there arises therisk that somewhere in the manufac-turing of your pump, a shortcut ortwo will be taken resulting in a pumpthat will not function adequately inthe field.

Also of interest to an engineer ismethod of payment. By tying pay-ment to achieved manufacturingmilestones, you can expedite themanufacturing of your pump, andthereby help to ensure on-time deliv-ery.

BUYING THE PUMP

Once you’ve clearly specified thepump, it’s time to place the order.This process can go smoothly, if you:• double check that the pump

you’re ordering is really the pump you want

• order the pump in time to allow for orderly manufacture

• have a post-award meeting, with-in one month of ordering, to ensure the order is clear and started

• don’t change the order unlesssafety or a major performanceproblem is involvedThe final three steps are self-ex-

planatory, but the first two deservesome explanation.

Double checking your order isespecially important for complexunits with an extensive specification.As a check, hold a pre-award meetingwith the manufacturer to clarify thebid. If your unit includes major auxil-iary equipment or systems, reviewand settle the basic unit plot plan atthis meeting.

Selecting a manufacturer can bedone in one of two ways: specify-and-evaluate or partnership-purchasing.Under the specify-and-evaluatemethod, you prepare a very detailedspecification and issue inquiries withextensive data requirements, thenthoroughly evaluate the data in theresulting bids and purchase based onthe numerical results of the evalua-tion. The evaluation generally takes

zontal end suction pumps for chemi-cal process and API-610, 7th edition,which addresses centrifugal pumpsfor petroleum refining). Some buyersuse these general specifications ver-batim, others use them as a base andadd supplement covering changesthey wish to incorporate.

A cardinal rule for any meaning-ful specification, whether home-grown or based on an industrystandard, is to avoid multiple tieredreferences to other specifications.With more than one tier of refer-ences, such specifications become toocomplicated to be meaningful. Forexample, when addressing govern-ment regulations, be sure to identifyand specify the exact rules and regu-lations the equipment has to meet.

Item Function Design Manufacture

1. Location and environment X2. Liquid pumped and properties X3. Hydraulic duty X4. Redundancy in pump arrangement X5. Future performance margin X6. Application margins X7. Type and number of pumps X8. Driver and arrangement X9. Minimum tolerable piping loads X10. Allowable seal leakage X11. Allowable noise X12. Minimum pump life X13. Mean seal life X14. Bearing life and basis X15. Mean period between overhauls X16. Materials of construction X17. Rotor design requirements X18. Hydraulic design requirements X19. Allowable stress X20. Type of shaft seal X21. Type of bearings and lubrication X22. Type of coupling X23. Type of base X24. Piping: systems required and construction X25. Auxiliary systems: specification X26. Instrumentation X27. Material tests X28. Welding procedures approval X29. Inspection during manufacture X30. Component and equipment tests X31. Painting and inhibiting X32. Documentation X

TABLE 1. ELEMENTS OF A PUMP SPECIFICATION TECHNICAL

Page 109: Centrifugal Pumps Handbook

110 The Pump Handbook Series

Engineers are not surprised by this;they know that technical endeavorsproceed best in a cooperative arrange-ment.

The specify-and-evaluate methodmight help you find a company thatwill furnish equipment nominallycapable of the same function for lessmoney (even when factoring in thecost of writing the specs and evaluat-ing the bids). But the issue isn’t justcost. The real value of partnership-purchasing is innovation in designand reliability of products.Partnership-purchasing is actually theway the pump industry used to oper-ate before competitive biddingbecame so popular. As an industry,we, the suppliers and purchasers,would do well to resurrect it. ■

J.T. McGuire is Director ofMarketing for the Huntington ParkOperations Division of the Ingersoll-Dresser Pump Company.

the form of a weighted matrix whichincludes:• energy consumption• maintenance cost• risk of lost production• purchase price• delivery

Manufacturers are free to bidwhichever pump they feel meetsyour specifications. That leaves you,the purchaser, to make the finaldetermination of whether a pumpmeets your requirements. Thus,you’ll need to build a rigorous inspec-tion regime into your selectionprocess. Over time, you can build alist of acceptable bidders to help nar-row the field.

Partnership-purchasing avoidsthe cost of preparing an elaboratespecification, issuing inquiries, andevaluating bids. Under this method,you select one manufacturer to workwith and provide just a minimal spec-ification. The manufacturer thenchooses the best pump for yourneeds.

To select the best manufacturerto work with you, you need to assessthe caliber of the various manufactur-ers that make the class of pumpyou’ve chosen. Your assessmentshould cover each manufacturer’s:• order engineering and manufac-

turing processes• emphasis on quality as an inher-

ent facet of all processes• product design philosophy• detail designs for and experience

with the class of pump requiredAfter choosing the best manufac-

turer to work with, you can negotiateprices for equipment according tosome fixed relationship to publishedprice lists. While you will incur somecosts in assessing manufacturers, thisprocess is likely to be less expensivefor a major product or a period of twoor three years between assessmentsthan open bidding would be.

Which approach is better? Forinnovation in design and reliability,I’ve found that partnership-purchas-ing yields distinctly better resultsthan specify-and-evaluate has.

Page 110: Centrifugal Pumps Handbook

The Pump Handbook Series 111

single look at Photo 1, above,should be enough to convinceanyone of the destructivenature of corrosion and abra-

sion on pumps, and lead to the ques-tion of how to prevent this fromhappening.

This article should serve as eithera primer or a reminder of factorsinvolved in properly selecting or trou-bleshooting a pump in corrosive and/orabrasive service. Historically, pumpselection has consisted of finding apump that will, “pump stuff from hereto there,” or that will “deliver so manygpm at such-and-such a head.” Agreater degree of sophistication leads to“and that will hold up in acid,” or “andthat will pump solids.” Obviously, themore that is known about the solutionbeing pumped, the more appropriatethe pump selection will be. An interre-lationship exists where the chemicaland physical properties of thepumpage determines the materials ofconstruction, which dictates pumpdesign, which affects pump perfor-mance, which in turn determines theproper pump selection. The more diffi-

cult selections involve services pump-ing corrosives and/or abrasives, andthese are the major factors governingwhich pump is chosen.

You don’t need to be a rocketscientist to select a pump, and youdon’t need a PhD in metallurgy tomake some basic materials selectionsand understand the reasoning behindthem. We all know that water will“rust” iron, acids “corrode” certainmaterials that come into contact withthem, and solids “wear” whenrubbed together; conversely, weknow that “stainless” steel is corro-sion resistant and that either a hardor soft “rubber” material will resistabrasion or wear. These simple factslead us to a closer examination of themechanisms of corrosive and abra-sive attack.

Corrosion is the wearing awayor deterioration of a material bychemical or electrolytic action orattack.

Abrasion is the wearing away ofa material caused by a solid rubbingor impinging on another. Abrasion

caused by the velocity of a liquid orgas is commonly called erosion.

Corrosion-abrasion is a combina-tion of both corrosion and abrasionthat results in an accelerated attackon material. It is generally moresevere than either corrosion or abra-sion alone, due to the severe wearcaused by the continuous abrasivedestruction of the passive protectivefilm built up by corrosion.

Table 1 shows the basic types ofcorrosion. Corrosion and abrasiontake many forms, and numerous com-binations of these forms exist.Detailed analysis of these combina-tions can be quite complex and goesbeyond the scope of this discussion.

MATERIALS

There is no material that willwithstand attack from all combina-tions of liquids and solids found inpumped solutions. However, a basicknowledge of material categories willgive us a general idea of what materi-als will and will not work in certainenvironments, and then we can zeroin on the right pump for a given job.

BY JOHN RINARD

A Common Sense Approach to Combating Corrosion and Abrasion

Photo 1. New and corroded centrifugal pump impellers.

A

CENTRIFUGAL PUMPS

HANDBOOK

Page 111: Centrifugal Pumps Handbook

112 The Pump Handbook Series

It becomes obvious with exami-nation of Table 2 that the mechanicalproperties of a material determine thedesign of a pump. Pumps constructedof hard materials are more difficult todesign (flanges, stack tolerances, andclearances), cast (sharp angles andcomplex shapes), and machine (drill,tap, and finish surfaces); non-metallicsmay need to be reinforced, supported,or protected with metal armor; andthin or highly stressed componentsmust be made of strong materials.

Figure 1 shows a typical configu-ration of both a chemical (corrosionresistant) pump and a slurry (abrasionresistant) pump. One can readily seethat the slurry pump’s hard metalmaterials of construction dictate theuse of through-bolt constructionrather than drilled and tapped holes.Less apparent are the facts that slurrypumps are generally more massivethan chemical pumps; are designedwith open clearances, blunt edges,and looser tolerances due to “as cast”hard metal surfaces and the need tohandle solids; and are commonlydesigned with metallic or nonmetallicliners. As a result of these design con-straints, slurry pump efficiencies suf-fer, and in most cases are lower thanchemical process pump efficiencies.

Identification of materials thatcan handle the liquid to be pumpeddoes not necessarily complete thematerial selection process; quite oftenthis step leads to other considera-tions. Options and compromisesalmost always present themselveswith either chemical or slurry pumpswhen comparing service life andwear with cost and availability.

CHEMICAL PUMPS

Wear. The chemical process industrygenerally considers that any corrosionrate equal to or less than 20 mils peryear is acceptable wear. This, howev-er, may be considered excessivedepending on either pump design(pump impellers with relatively thinvanes and shrouds effectively seedouble this wear rate because theyare totally immersed in the liquid andtherefore exposed to attack from bothsides) or a need for extended servicelife for pumps in critical services andinaccessible or remote locations.

FIGURE 1

Pump design comparison. On the top is a hard iron slurry pumpwith side suction. The bottom is a stainless ANSI B73.1 chemicalpump with an expeller-type seal.

Page 112: Centrifugal Pumps Handbook

The Pump Handbook Series 113

and recirculation result in increasedliquid and solids contact with wettedpump surfaces, as well as unpre-dictable angles of impingement.Proper pump selection, therefore, dic-tates selection at or near the best effi-ciency point of the pump. Selectionjust to the left of best efficiency isconsidered good practice, as illustrat-ed by Figure 2.

Analysis can often trace the causeof pump problems to operating thepump at or too close to shut off (to thefar left of best efficiency) because thepump is oversized. Intentional over-sizing may occur through the use ofsystem design safety factors, selectionfor future increases in performance,and using an existing pump withoutconsideration of size. Unintentionaloversizing may occur because of mis-calculations or changes over time inthe process or the piping system. Theend result is the same; the pump isoperating too far away from the bestefficiency point.

When analyzing pump perfor-mance, we must think in terms of apumping system rather than just thepump. A system consists of the pumpand all the related piping, valves, andprocess equipment on both the suc-tion and discharge sides of the pump.All of these items directly affect thepump performance in that the “sys-tem curve” (which can be analyticallyderived from the pressure drop/resis-

Cost. Some material costsmay be prohibitively highand therefore lead toselection of less corrosionresistant alternatives or alined rather than a solidmaterial pump. Availability. While mate-rials such as 316 stainlesssteel, CD4 MCu, andAlloy 20 are commonlystocked and available forchemical pumps, alloyssuch as monel andHastelloy are more likelyto be special orders.“Standard” materials ofconstruction vary frommanufacturer to manu-facturer. Depending onthe pump type and themanufacturer, materialavailability can varyanywhere from being instock to needing up to several monthslead time.

Lined or coated pumps and non-metallics offer possible solutions to thehigh cost and long lead times of non-stock special metallic materials.Lining pumps is where nonmetallicsreally shine, and they are less expen-sive and more readily available thanspecial metallics. However, just asthere is no single metallic that is goodfor handling every solution pumped,there is also no single nonmetallic forall services. Each must be carefullyselected to fit the service.

SLURRY PUMPSThe selection of abrasion-resis-

tant materials for slurry pumps,much the same as corrosion-resistantchemical pump material selection,also involves consideration of servicelife (wear), cost, and availability.Abrasion, unlike corrosion, is general-ly combated by the use of either veryhard materials or soft, resilient elas-tomeric materials. Hard materials aregenerally used for slurries with largeor sharp solids. Soft, resilient elas-tomeric materials are used for smallor blunt solids. Once again, here wefind that non-metallic elastomers lendthemselves for use as pump liners.

CHEMICAL-SLURRY PUMPSIt was mentioned earlier that a

pumped solution that is both corro-

sive and abrasive, an acid sludge, forexample, presents the greatest chal-lenge in pump selection. Many materi-als are essentially suitable to eithercorrosion or abrasion, but not to both;titanium, for example, is a very strong,corrosion-resistant material, but it isunsuitable for slurries because of itssoftness, and white iron is a very hard,abrasion-resistant material that is notpractical for corrosive conditions. Non-metallic elastomers, on the otherhand, may be used in a service that isboth corrosive and abrasive. Whenselecting elastomers, considerationmust be given to solids size and config-uration, temperature, and a pumpdesign that must generally precludeliquid contact with any metallic armoror reinforcing.

PUMP PERFORMANCE

The efficiency of a pump as wellas the location of the operating pointon the pump performance curve isoften overlooked or ignored duringpump selection. The location of theoperating point is overlooked moreoften than the pump’s efficiency.This alone will contribute as much asany other factor to pump failurewhen abrasion or corrosion-abrasionare present. Efficiency is a measure-ment of smooth flow—and thereforereduced turbulence and recircula-tion—within the pump. Turbulence

Type Characteristics RemarksGeneral Uniform attack over entire exposed surface Most common type of corrosionErosion-Corrosion Corrosion accelerated by erosive action

of fluid or slurry vortexCrevice Localized attack at crevices or Commonly found at gasketed or flanged

stagnant areas surfacesGalvanic Occurs when two dissimilar metals are

immersed in a corrosive or conductivesolution

Intergranular Grain boundary attack Weld decay is a type of intergranular corrosion occurring in areas adjacentto a weld

Cavitiation Pitting on high pressure areas such asimpeller vane tips and/or low pressure areas such as eye of impeller vanes or trailing edge of impeller vanes

Pitting Localized accelerating attack by chlorides; Common in 304 and 316 SS, and A20associated with stagnant conditions

Selective Leaching Dissolves one component of an alloy Zinc removed from brass or bronze iscalled dezincification; when grey cast ironis attacked, graphite is left undisturbed

TABLE 1. TYPES OF CORROSION

Page 113: Centrifugal Pumps Handbook

114 The Pump Handbook Series

tance to flow across various in-linehardware) dictates where the pumpwill operate on its curve, the pumppoint of rating. Less sophisticatedconsiderations include rules of thumbsuch as:

• Keep the suc-tion piping asshort andstraight aspossible.

• Slope the suc-tion pipingtoward thepump suctionwhen han-dling slurries.

Centr i fugalpumps tend tobecome unstablethe closer theyapproach eithershut off (zero flow)or maximum flow.This instabilitymay be manifest-ed in cavitation,recirculation, andturbulence. Recir-culation and tur-bulence can resultin a liquid temper-ature rise in thepump that cancause acceleratedcorrosion as wellas erosion-corro-sion.

PUMP SELECTION

The final selection decision ismade by the pump user. This decisionmay be more subjective than analyti-cal, but should include such factors as:

• Availability (of both pump andparts)

• Maintainability• Reliability• Service life• Standardization• Cost

There are always trade offs. Theuser ultimately makes a selectionbased on the priorities that best meetthe process needs. This paper haspresented an overview of corrosionand abrasion factors that should be apart of that selection process.

OPTIMUMSELECTION

BEST EFFICIENCYPOINTS

FIGURE 2

Optimum pump selectionresults in a pumpoperating just to the left of itsbest efficiency point.

Typical Mechanical PropertiesCategory Subcategory Material Typical Tensile Elongation

Hardness* Strength (Min % in 2")(Min, psi)

Metallics Ferrous Steel 150 Brinell 70,000 22Ductile Cast Iron 160 Brinell 60,000 1827% Chrome 600 Brinell 80,000 Nil

Stainless 304 SS 150 Brinell 70,000 35316 SS 150 Brinell 70,000 30CD4MCu 225 Brinell 100,000 16A20 125 Brinell 60,000 35Hastelloy B/C 225 Brinell 75,000 20-25

Copper base Brass 60 Brinell 37,000 30Bronze 65 Brinell 35,000 18

Miscellaneous Aluminum 130 Brinell 65,000 8Titanium (pure) 200 Brinell 80,000 18Zirconium (pure) 210 Brinell 55,000 12

Non-Metallics Elastomers Rubber (gum) 35 Durometer A 3,500 500Neoprene 55 Durometer A 3,000 650-850Urethane 75-95 Durometer A 4,500-7,500 250-900

Plastics Teflon (PTFE) 50-65 Shore D 3,000-4,000 200-400Epoxy (cast) M75-110 Rockwell 2,000-12,000 NilPolypropylene R85-95 Rockwell 5,000 500-700

Ceramics Silicon Carbide 2,500 Knoop 44,500 NilAluminum Oxide 1,000-1,500 Knoop 22,000-45,000 Nil

TABLE 2. MATERIALS COMPARISON

There are numerous alloys, formulations, and compounds of metallics and non-metallics; those shown are typical andare not to be considered all-inclusive.*Conversion relationships of hardness scales/numbers are discussed in ASTM E140 and Metalcaster's Reference andGuide, 2nd Edition, 1989, The American Foundrymen's Society, Inc., (for metals), and ASTM D2000 (for rubber).

ACKNOWLEDGMENT

My thanks to Dr. GeorgeCalboreanu, Chief Metallurgist,Western Foundry, a division of A.R.Wilfley and Sons, for his materialsexpertise and assistance in preparingthis paper. ■

John W. Rinard holds a bachelor’sdegree in industrial engineering fromTexas A&M University. His experienceincludes positions in Sales Engineeringand Management with the BuffaloForge Company and the DurironCompany. He is presently with A.R.Wilfley and Sons.

Head

Flow

Page 114: Centrifugal Pumps Handbook

The Pump Handbook Series 115

CENTRIFUGAL PUMPS

HANDBOOK

LOCATIONA vertical turbine, mixed flow

or axial flow pump’s location in asump is critical to good perfor-mance. Figures 1 and 2 providegood design criteria for sump lay-out. These criteria are based on amaximum bell entrance velocityof 6 ft/s. However, because belldiameters vary from manufacturerto manufacturer, these ratios mustbe adjusted to accommodate thedifferences.

According to the U.S. ArmyCorps of Engineer’s design guide,”For satisfactory pump perfor-mance based on research and pro-totype experience, recommendedsubmergence, S, should be 1.25 Dor greater, and the dimensionlessflow ratio through the individualpump should not exceed a valueof 0.40 for:

Q/ √gD5

where

Q= discharge, cfs

D= pump bell diameter, ft

g= acceleration due to gravity, 32.2 ft/s2

Submergences that are lessthan, and flow rates that exceedthe above limits were investigated,and more complex designs wererequired for satisfactory hydraulicperformance.”

The recommendation of 1.25D minimum submergence is suit-able for storm water and floodcontrol pumps (provided a vortexsupressor beam is used as illus-trated by Figure 2); however, forcontinuous service pumps a sub-mergence of 1.75 D is recom-mended. If the submergence isless than these values, the belldiameter must be enlarged. Forinstance, to meet a 1.25 D sub-mergence value, the bell diame-ter should be enlarged to produce

positive suction head (NPSH), thenoise created by a vortex comesand goes as the vortex comes andgoes. To mitigate submerged vor-tex formation, apply the followingstrategies:• Place a cone under the bell.

• Employ splitters.

• Fill-in intake corners.

• Use diffuser screens.

HIGH VELOCITYAs a rule, high velocity to a

pump in the intake and/or at thebell leads to reduced life of thepump. For a given head and capac-ity, today’s pumps operate atapproximately double the speed ofthe pumps in use before the

Recommendations For Vertical Pump IntakesBY: HERMAN GREUTINK

an average entrance velocity of 3.3ft/sec. This velocity may be a bitconservative, but the cost of enlarg-ing the diameter is low and the bene-fits are tangible.

VORTICESIf a vortex still occurs after you

have followed the above guidelines, itis not generally difficult to alleviate. Ittakes very little energy to form a vor-tex; therefore, it takes very little ener-gy to get rid of it!

Submerged vortices, however,can be troublesome. These vorticeswill touch the floor and/or wall of anintake. They are the result of swirlingmasses of water next to or under thepump and are not continuous.Although submerged vortices soundlike cavitation due to the lack of net

1 D 2 D 2 D 1 D

Velocity preferred to be1 ft/sec

.5 DD

FIGURE 1

.8 D

Page 115: Centrifugal Pumps Handbook

116 The Pump Handbook Series

1960’s. The net result of these higherspeeds is a drastically increased fre-quency of pump repairs. Slowingdown continuous service pumpspeeds may be more expensive ini-tially but the long-run savings onmaintenance will more than compen-sate for the increased pump costs.

A high velocity stream aimed ator near the pump could also be asource of premature failure. A fluidforce of this nature should be dif-fused by piling, screens or walls infront of the conduit outlet. Figure 3provides a simplified depiction of dis-tances required to diffuse a highvelocity flow out of a conduit.

The breakup of jet streams canbe achieved by baffles as shown inFigure 4. This configuration also pro-motes better distribution to multiplepumps.

DIVIDING WALLSBecause short dividing walls are

not recommended, they are not pic-tured in any of the figures. (Figure 1shows no dividing wall while Figures2 and 4 show long dividing walls.)With multiple pump stations, thefront of the short walls can propagatevortices when one or more pumpsare out of service. So it is better tohave no walls than short walls. Longwalls provide easy support for thepumps, as well as drainage for indi-vidual pump sumps when stop logsare used.

INTAKE TESTSWhen guidelines such as those

published by the Hydraulic Instituteand the British Hydro-mechanicsResearch Association (BHRA) cannotbe followed, model intake testsshould be performed, especially forpumps larger than 50,000 gpm. ■

A

A

1D1D

(TYP)

0.25D

2D

R=2D135°

Vorte

xSu

ppre

ssor

Beam

Pum

p Ba

y

Divi

der W

alls

Rounded2D

135°

6D

45°(wing)WallPLAN

S = SubmergenceD = Pump Bell Diameter

0.25D

1.5D

0.5D

1.0D

1.25

D

S

0.5D1.0D

MinimumWaterLevel

Curved (wing) Wall

Section A-A

FIGURE 2

W

Page 116: Centrifugal Pumps Handbook

The Pump Handbook Series 117

Trashrack

Flow

Baffles Forebay Pump Bay

SuctionBells

Plan

SuctionBells

D = Suction bell diameter, ft.

D

FIGURE 4

30d

10d4-5d

d V

III

III

IV

x

Vm

Area I - Potential CoreArea II - Transition

Area III - Similar Velocity ProfilesArea IV - Jet Center Line Wanders

Up to about 30 times diameter d, the formula

= 6.5 x is used to determine Vm

V and Vm (ft/sec.), d and x (ft.)

Vm___V

d___x

FIGURE 3

U.S. Army Corps of Engineer’s Design Instructions for FloodControl Pumps

REFERENCES1. U.S. Army Corps of Engineers,

Engineering Technical LetterNo. 1110-2-313 ”HydraulicDesign Guidance for Rectan-gular Sumps of Small PumpingStations with Vertical Pumpsand Ponded Approaches.”

2. Prosser, M. J. ”The HydraulicDesign of Pump Sumps andIntakes.” British Hydromech-anics Research Association.

3. Hydraulic Institute Standards,14th Edition

Herman Greutink is vice presi-dent and technical director forJohnston Pump Company inBrookshire, TX.

Page 117: Centrifugal Pumps Handbook

118 The Pump Handbook Series

CENTRIFUGAL PUMPS

HANDBOOK

ydraulic excitation forces andpressure pulsations createdby excessive flow decelera-tion at partial load have a pro-

found impact on the possible failureof a variety of pump components.These forces and pulsations are theresult of flow recirculation in theimpeller inlet, diffuser or volute sec-tions of the pump. Some degree ofrecirculation is present in every cen-trifugal pump below a specific flowrate representing the ”onset of recir-culation.” In fact, recirculation is ofminor concern for the majority ofpump designs. On the other hand,excessive recirculation can beextremely harmful and destructive.Consequently, the onset of damagingrecirculation is of greater concern topump operators than the onset ofrecirculation itself.

IMPELLER INLET RECIRCULATIONThree physical mechanisms trig-

ger flow recirculation during partialload at the impeller inlet: 1. deceleration of the velocity

upstream to the impeller relativeto the velocity in the impellerthroat

2. pressure gradients perpendicu-

lar to the direction of mainthrough-flow

3. excessive incidence (i.e., differ-ence between blade angle andflow angle at the impeller vaneleading edges).

The primary geometrical para-meters impacting the above phenom-ena are:• impeller throat area

• angle of approaching flow

• impeller vane angles

• ratio of impeller eye diameter tohub diameter

• ratio of vane tip diameter to hubdiameter

• impeller shroud curvature

• impeller leading edge position (inplanar view and meridional sec-tion).

However, no simple generalrelationship exists between the onsetor amount of recirculation and thegeometry of the impeller. Relation-ships have been derived that arevalid only for particular families ofimpellers (Ref. 1). Applying theserelationships to impellers developed

according to design rules differingfrom those underlying the correlationwould be very misleading.

IMPELLER OUTLET RECIRCULATIONDownstream to the impeller the

flow may be decelerated in a station-ary component of the casing. Thisdeceleration can occur in a diffuser, avolute, an annular casing or a combi-nation thereof. The physical mecha-nisms of downstream decelerationare quite similar to those previouslymentioned for upstream recircula-tion. They are:

1. deceleration of the absolutevelocity from the impeller outletto the throat of the casing

2. incidence at the diffuser vanes orvolute cutwater

3. pressure gradients perpendicularto the direction of the main flow(particularly for semi-axial oraxial pumps).

The main geometrical parame-ters impacting flow recirculation atthe outlet are:

• velocity distribution at theimpeller outlet as determined bythe geometry of the impeller

HydraulicInstabilities and

CavitationCauses, Effects and Solutions

BY: J.F. GUELICH AND T.H. MCCLOSKEY

H

Page 118: Centrifugal Pumps Handbook

The Pump Handbook Series 119

TABLE 1. MEANS TO UNDERSTAND AND TO MODIFY THE SHAPE OF THE HEAD/CAPACITY CHARACTERISTIC (HCC)Symptoms

Q-H-curve Internal pressures Axial thrust, β/ω = f (Q)

Possible causesormechanisms

Possible remedies

NOTE: any modifications may have side-effects which should be carefully assessed

• Insufficient recirculationat impeller inlet(insufficient centrifugalhead increaseat low flow)

• Insufficient recirculationat impeller outlet(insufficient exchangeof momentum betweenimpeller anddiffuser at low flow)

• Excessive recirculationat impeller inlet

• Excessive recirculationat impeller outlet

1. HCC droopingtowards shut-off

2. Excessive shut-offhead and/or excessive shut-offpower

Typical for nq < 30

Typical for nq > 35

H = total dynamic headQ = net flow through

pump

ψp = static head rise ofimpeller / (u2

2/2g)Hc = head rise in casing

Fax = axial thrust towardssuctionβ

ω

• advance impeller vaneleading edge (reduce d1 eff)

• advance return vanetrailing edge (reduce cou)

• increase d1a (increase ∆ d1 eff)• reduce hub dia. (increase ∆ d1 eff)

• reduce a3 or s8• increase b3 / b2• reduce gap A, increase overlap• increase δTE• increase b2• reduce gap B (with small nq) but be-ware of increased pressure pulsations

• reduce d1i (increase d1 eff)• cut back return vane (increase cou)

• reduce d1a• increase hub dia.• inlet ring } reduce

d1 eff

• increase a3 or s8 • reduce b3 / b2• reduce b2• increase δTE

Rise of ψp

towards Q=0

Hc, flat or droopingtowards Q=0

Hp, ψρ flat towardsQ=0

Rise of Hctowards Q=0

3. HCC with saddleor flat position

Axial thrust excursions

S = outlet recirculation onshroud

B = outlet recirculation onhub

B

Steep rise of NPSH littleor not at all affected byimpeller inlet

• Flow separation in diffu-ser (volute) but not yetfully developed recircul-ation

• Shifting of flow patterns(zones of recirculation /flow separation)

• Very sensitive to manu-facturing tolerances ofdiffuser and impeller

• Very sensitive to impellerinlet flow conditions andimpeller inlet geometricparameters

• Sensitive to axial rotorposition: HCC and Fax

• Excessive flow accelera-tion in diffuser or volutethroat

• Cavitation in diffuser orvolute

• Axial thrust fluctuations and sen-sitivity to rotor position / axialstage stacking tolerances can beeliminated by reducing gap Aand introducing proper overlap

• To remove saddle type instabilitydetailed flow analysis and / or testing often required

• Differences in stage geometry toget onset of flow separation indifferent stages at different flow rates

• Increase diffuser or volute throatarea

0.5

0.5

= ratio fluid/shroud rotation

4. HCC too steep athigh flow rates

Sudden decrease of β dueto outlet recirculation

ψp

Q

Hc

Q

β

Q

Q

Q

H

ψp

0.5

β

Q

Hc

Q

Q

Hc

FaxH

SQ

S

H

H

Q

Q

30< nq < 60

nq < 60

H

Q

η

QP

Q

H

Q

Hc

Q

Hp

NPSH3%

Q

δTE

δTE

Page 119: Centrifugal Pumps Handbook

120 The Pump Handbook Series

itself as well as the velocity dis-tribution at the impeller inlet

• diffuser or volute throat area

• diffuser vane or cutwater angles

• ratio of impeller outer diameterat tip to hub (oblique cut of radi-al, semi-axial or axial impellers).

INCREASING SHUT-OFF HEADAmple evidence suggests that

increasing the recirculation at theimpeller inlet and/or outlet increasesthe head. A number of geometricalparameters can be altered to increasethe head in this manner. If the flowversus pressure head (Q-H) curve isdrooping towards shut-off, invokingrecirculation may be appropriate,especially since the shut-off head isparticularly affected by an increase inrecirculation.

Table 1 presents typical shapesof Q-H curves, explains the physicalmechanisms responsible for thesecurve shapes and suggests possibleremedies and geometric parametersby which the shape of the Q-Hcurve can be corrected. However,the reader should be aware thatthese remedies may produce unde-sirable side effects (e.g., reduction inhead or efficiency at best efficiencypoint (BEP)).

DAMAGING RECIRCULATIONFor every type of pump there

exists a range of optimum recirculat-ing flow, and operating a pump with-in it avoids the risk of unstableQ-H-curves on one side and the riskof damaging levels of recirculation onthe other. This range is illustratedqualitatively by the Figure 1 graph.Unfortunately, there is no establishedmethod to predict exactly the onset ofdamaging recirculation. An unaccept-able level of recirculation can bedetermined indirectly, however, inindividual cases by applying the fol-lowing strategies:

• Measure cavitation noise toassess the risk of cavitation ero-sion.

• Test for vibration.

• Monitor shut-off head or shut-offpower. If excessive, this mayindicate inordinate recirculation.

• Measure the radial or axialhydraulic excitation forces. Asudden rise in these forces as theflow rate is reduced and/or con-sistently excessive levels of theseforces could indicate an unac-ceptable degree of recirculation.

• Measure pressure pulsations. Asudden rise in pressure pulsa-

tions as the flow rate is reducedand/or consistently excessivepulsation levels could indicatedamaging flow recirculation.However, a sudden rise of pres-sure pulsations might also resultfrom standing wave resonance.For this reason, to avoid the pos-sibility of misinterpretation ofhigh loads of pressure pulsation,careful testing and data analysisis imperative when diagnosingthe true nature of pulsation.

Hydraulic excitation forces andpressure pulsations may be responsi-ble for a number of possible compo-nent failures. Table 2 details the rootcauses and mechanisms of such fail-ures along with possible remedies.

Partial load flow phenomena alsostrongly influence pump vibration.Observed vibration phenomena caus-es and mechanisms, along with possi-ble remedies, are given in Table 3.

CAVITATION EROSIONIf, as a result of recirculation, the

local pressure at the impeller inletdrops below the saturation pressureof the pumped liquid, vapor bubblesare generated and are then swept bythe flow into zones of higher pres-sure, where they implode and maycause erosion of the impeller.

To eliminate or reduce cavitationdamage, the following remedies areavailable:• Change operation procedures if

damage occured at partial load oroverload.

• Increase net positive suctionhead available (NPSHA).

• Reduce speed or use a variblespeed drive if partial load isrequired.

• Increase cavitation resistance ofmaterial.

• Modify geometry of impeller(profiling of blades, impeller re-design).

• Improve inlet flow conditions bygeometric modifications.

• Increase gas contents.

In addition, Table 4 outlines cavi-tation damage mechanisms and offerscorrelating remedies. As illustrated

Optimum Design

HMax-Ho

Lcav

H=f(Q)unstable

H=f(Q) stable

Pressurepulsations /noise

Hydraulicexcitationforces

Low freq.pulsations

QRec

CNL

FIGURE 1. OPTIMUM AMOUNT OF CIRCULATION

Page 120: Centrifugal Pumps Handbook

The Pump Handbook Series 121

TABLE 2. EFFECT OF HYDRAULIC EXCITATION ON COMPONENT FAILURE

1. Fracture of impeller blades atoutlet, diffuser vanes at inlet, tiebolts, instrument piping, or other components

• High dynamic stresses inducedby pressure pulsations (im-pingement of wake flow from impeller blade trailing edge ondiffuser vanes or volute cut-water)

2. Side plate breakage

3. Mechanical seals

• High dynamic stresses inducedby pressure pulsations

• Impeller side plate resonance if z3 - z2 = 2 and z3 n/60 closeto impeller side plate naturalfrequency

• High pressure pulsations causedby wake flow or recirculation /separation

• High frequency pressure pulsa-tion due to cavitation

• Shaft vibrations

• Increase gap B by cutting back(1) diffuser vanes if diffuser throat

does not increase by more than 3%

(2) impeller blade trailing edge (head of pump will be reduced unless speed cannot be adapted)

• Reduce excitation at part load bymodifying hydraulic components(careful analysis and redesign)

• Increase gap B (see previousitem)

• Change z3 / z2 combination• Modify natural frequency• Reduce exciation at part load by

modifying hydraulic components

• (see above)• Reduce cavity volume by rede-

sign of impeller and / or inlet• see table 3

• There are a number of other failure mechanisms related to design, material selection and quality

RemarkPressure pulsations and dynamicstresses are expected to decreasewith a power of -0.77 of gap B.For example to achieve half of theoriginal level gap B must be increased by a factor of about 2.5

• Insufficient quality of impellercasting and / or finish (notcheffect)

• Insufficient thickness of impellerside plates

• There are a number of other failure mechanisms related todesign, material selection andquality

Failure / incident Possible hydraulic causesor mechanisms

Possible remedies Possible non-hydrauliccauses / remarks

4. Excessive labyrinth wear

5. Failure of radial bearings

6. Failure of axial bearings

• Excessive radial thrust

• Excessive vibration

• Excessive radial thrust• Excessive vibration

• Axial thrust excursions• Excessive labyrinth wear (high

leakage increases rotation onshroud; reduces rotation on hubwith multistage-pumps)

• Reduce flow asymmetries around impeller by- double volute in case of single

volutes- analyzing / eliminating cause

of asymmetry (casting tolerances, differences inresistance in channels ofdouble volutes, discharge andsuction nozzle,...)

• see table 3

• see above item• see table 3

• see table 1, item 3• Replace wear rings

• Thermal deformations of casingand rotor

• Mechanical / design

• Mechanical / design• Transients

at partial load on the pressure side ofthe blades, flow recirculation wasidentified as the most probable cause.

To improve the partial load rangeand thereby increase the impellerlife, an inlet ring was designed andinstalled in the pump. Figures 2 and 3show the fluid-borne and solid-bornenoise prior to and after this modifica-tion. Prior to modification the noiserecorded at 100% and 80% flow isvirtually equal. Since no erosionoccurred during more than 50,000hours of operation at full load, thisevidence suggests that the operationat 60% load was entirely responsiblefor the damage. As illustrated by thefigures, the modification of the pump

by the following case study, geomet-ric modification of the impeller is fre-quently the only feasible solution.

CASE HISTORYAfter a boiler feed pump had

operated for more than 50,000 hourswith no trace of cavitation on the suc-tion impellers, the load demand ofthe process changed, requiring pro-longed partial load operation. Thepump operated about 1000 hours at60% load and 1100 hours at 80% loadbefore cavitation damage was discov-ered on the pressure side of theimpeller blades. The attack variedbetween 2-4mm from blade to blade.Since the cavitation damage occurred

decreased the noise at 60% flow tothe unmodified 100% flow noiselevel, and the erosion problems weresolved. ■

ACKNOWLEDGEMENTS This article summarizes the

results of investigations on hydraulicinstabilities and cavitation erosionsponsored by the Electric PowerResearch Institute (EPRI), Palo Alto,CA and conducted under EPRI RP 1884-10. The authors are gratefulto R. Egger, W. Handloser and A.Roesch, who carried out the exten-sive test program.

Page 121: Centrifugal Pumps Handbook

122 The Pump Handbook Series

REFERENCES1. W.H. Fraser. ”Recirculation in

Centrifugal Pumps.“ ASMEWinter Annual Meeting. 1981.

2. J.F. Guelich et. al. ”Feed PumpOperation and Design Guidelines.“EPRI Final Report TR-102102.June, 1993.

J. F. Guelich is manager ofhydraulic pump design for Sulzer PumpDivision in Winterthur, Switzerland.

T. H. McCloskey is manager ofturbo-machinery at the Electric PowerResearch Institute in Palo Alto, CA anda member of Pumps and Systems

Editorial Advisory Board. He is a fellowof ASME and a member of theHydraulic Institute’s technical commit-tee on pump intakes.

NOMENCLATURE

A amplitute

a3 diffuser throat width

b2 impeller exit width

b3 diffuser inlet width

Com absolute velocity at meridonal inlet point

CNL cavitation noise level

cou absolute velocity upstreamof impeller

D1 impeller eye diameter

D2 impeller outer diameter

d1 D1/D2

d1eff impeller vane inlet diameterwhere flow enters theimpeller

Fax axial thrust towards suction

f frequency

fn rotational frequency

H head per stage of pump

Hc head rise in casing

Hp static head rise of impeller

Lcav cavity length

nq pump specific speed (metricconvention)

Q flow rate

QSL flow at shockless entry

S outlet recirculation onshroud

S8 volute throat area

u1 circumferential velocity

z2 number of impeller vanes

z3 number of diffuser vanes

ß angular velocity of liquid

δTE trailing edge angle

σ slip factor

σu1 cavitation coefficient(2gNPSH/(u1)

2)

ψp static pressure rise ofimpeller

ω angular velocity of impeller

Subscripts:

av available

BEP best efficiency point

rec recirculating flow

X

X XX X X

X X

X

original

modified

N/m2

100,000

80,000

60,000

40,000

20,000

00 10 20 30 40 50 60 70 80 90 100 110 % Flow

NL

X

X X X XX

X XX X

0 10 20 30 40 50 60 70 80 90 100 110 % Flow

120

100

80

60

40

20

0

modified

original

CVm/s2

FIGURE 2. FLUID-BORNE NOISE IN A FEED PUMP (1 TO 180 kHz)

FIGURE 3. SOLID-BORNE NOISE AT A FEED PUMP (10 TO 180 kHz)

Page 122: Centrifugal Pumps Handbook

The Pump Handbook Series 123

TABLE 3. INTERACTION BETWEEN FLOW PHENOMENA AND VIBRATIONObserved vibration Possible hydraulic causes

and mechanismsPossible remedies Major non-hydraulic

causes / remarks

1. Subsynchronous peakclose to fn

• Increased labyrinth preswirl ⇒reduced rotor damping

• Unloading of bearings due to change in radial thrust withflow

• Rotating stall• Labyrinth wear

⇒ increased leakage ⇒ increased preswirl⇒ reduced rotor damping

Spectral component

2. Synchronous vibration

3. Supersynchronous peaks

A = amplitudef = frequencyfn = rotational frequencyz2 = number of impeller

blades

< 1.0

> 1.0

Q/QBEP

all

all

all

• Hydraulic unbalance due tovarious impeller tolerances

• Pressure pulsations caused bywakes from the impeller bladetrailing edge

• Harmonics other than bladepassing frequency are due toimpeller casting tolerances(pitch of the blades)

• z3 - z2 = +/- 1 resulting innon-zero radial blade forcecomponent at z2 fn

• Increase rotor stiffness anddamping by introduction of plain labyrinths or shallow serrations only

• Reduce preswirl to labyrinths by swirl brake

• Impeller or diffuser redesign

• Reduce casting / manufacturingtolerances of impeller (precisioncasting, ceramic core proce-dures, manufacturing) and implement more stringent question/answer procedures

• Peaks nearly always present.If excessive:- Increase gap B (see table 2,item 1)

• Harmonics other than bladepassing frequency: reduce impeller casting tolerances

• Change number of impeller ordiffuser vanes

• Reduce excitation force by properstaggering of impellers on shaft

• Labyrinth design• Thrust balancing device does

not provide sufficient rotor damping

• Bearing design / bearing unloaded

Remark:Instability may be recognized byvery steep increase in amplitudewith increasing speed

4. Broad band shaft vibrations

5. Structural resonances below frequency of shaftrotation excited by broadband hydraulic forces(e.g. bearing housing, bed plates, piping, ...)

6. Rotating stall

7. Surge-like strongpulsations

typicallybelow50% ofBEP flow

typicallybelow90% ofBEP flow

low

• Flow recirculation at impellerinlet and outlet

• Some broad band vibrationsare unavoidable. If excitation isexcessive this can be due tooversized throat areas of diffuser or volute, oversizedimpeller eye or excessiveincidence

• Fluctuating cavities

• Stall cells in diffuser or impellerrotating with a frequency belowthe frequency of rotation

• Vapour core forming in the suction pipe at low NPSHav dueto strong part load recirculation

• If excessive: reduce diffuser orvolute throat area; reduce impeller eye (careful review ofhydraulic design required)

• As a cure of the symptoms therotor damping might be increased (swirl brakes, labyrinth redesign, see above)

• Reduce cavitation extension(higher NPSHav, redesign ofimpeller or inlet)

• Analysis and redesign ofhydraulic components

• Increase rotor damping (swirlbrakes, labyrinth redesign)

• Structures upstream of im-peller to avoid formation of core(flow straightener, cross, inlet ring, hub diameter, “back-flowcatcher”)

• Impeller / inducer redesign• Air admission (if possible)

Remark:It is typical that structural reson-ances excited by broad bandforces do not depend on the speed of the shaft

Remark:The peaks are expected to beproportional to the rotor speed

Subsynchronous vibrationsincrease with time

A

f

fnInstability

Afn

f

Afn Z2fn 2Z2fn

f

Afn

f

A fn

f

Page 123: Centrifugal Pumps Handbook

124 The Pump Handbook Series

TABLE 4. CAVITATION DAMAGE MECHANISMS AND REMEDIES

Type of cavitation /damage pattern

Flow mechanisms likely to inducedamage

Possible causes Possible remedies

1. Suction side of blade,starting close to leadingedge of blade

Sheet cavitation on suction side ofblade at Q < Q SL

• Damage near shroud• Damage near hub

2. Suction side of blade,within channel

Sometimes erosion is alsoobserved on the shroud and /or the hub or on pressure sideof the blade

Vortex cavitation on suction side ofblade at Q < Q SL at low σ. Vortexesdeveloping downstream of a long,thick cavity attached to the blades.Bubbles created in the vortexes areswept away by pulsating flow and can implode anywhere in thechannel.

QSL = flow at shockless entry

• High incidence• Unfavorable leading edge profile

• Outer blade angle β1a too large• Inner blade angle β1i too large

• High incidence at low σ (typicallyσu1 av = 0.15 to 0.3)

• Insufficient NPSHav• Insufficient cavitation resistance

of material

• Increase flow rate• Reduce blade inlet angles• Improve leading edge profile• Reduce incidence by inlet ring• Reduce impeller eye diameter if

above optimum range• Increase pre-rotation

• Reduce impeller eye diameter ifabove optimum

• Reduce blade inlet angle• Increase NPSH available• Increase cavitation resistance of

material

3. Pressure side of blade, anylocation, starting close toleading edge of blade

4. Pressure side of blade,damage at outer half ofimpeller width startingclose to leading edge

5. Pressure side of blade,damage near hub. Difficult todistinguish from item ③ un-less it can be determinedwhether pump has operatedat partload or above BEP

6. Damage on hub or shroudor in fillet radii

Sheet cavitation on pressure side atQ > Q SL

Bubbles in free stream generated byshear flow due to partload recircula-tion. Bubbles impinge on pressureside of vane.

Excessive partload recirculation crea-tes a negative incidence near hub

ω0 Qmin actual

Comer vortex cavitation often com-bined with high incidence

• Negative incidence due to excessive flow

• Unfavorable leading edge profile• Excessive (run-out) flow

• Excessive flow deceleration(excessive impeller eye diameter,excessive impeller throat area,excessive blade angles)

• Negative incidence near hub dueto partload recirculation

• Blade angles not properlymatched to the flow

• Fillet radii too large

• Reduce flow rate• Increase blade inlet angle

(but beware of partload cavitation)

• Improve leading edge profile ifdamage close to leading edge

• Increase flow rate• Impeller redesign (decrease eye

diameter / throat area / bladeangles)

• Inlet ring at the impeller entrance(reduce deceleration, reduceshear flow effects)

• Increase blade angle at hub• Reduce recirculation• Improve leading edge profile to

reduce flow separation near hubunder recirculation

• Reduce preswirl (vanes, ribs,backflow catching elements) ofrecirculating flow

• Impeller redesign (adaption ofblade angles)

• Required fillet radii

ω0 QBEP

ω0 Qmin theoretical

Lcav

U1

COM

Page 124: Centrifugal Pumps Handbook

The Pump Handbook Series 125

CENTRIFUGAL PUMPS

HANDBOOK

he September 1994 issue ofPumps and Systems featured anarticle entitled ”Low FlowOptions.” Among other things,

it discussed the single port diffuserpump design credited to Dr. U. M.Barske. A high-speed version of thisdesign was popularized in 1959 withits application as a water injectionpump for the Boeing 707 airliner. Anindustrial version of this concept wasfirst marketed in 1962 and tens ofthousands have now been placedworldwide. The single port, or partialemission, pump has been developedcommercially by a number of manu-facturers. The designs accommodateboth two and four pole electric motordrives and are most often applied forlow specific speed services. Figure 1shows the basic geometry of an openimpeller, ”partial emission” diffuser,design. The high-speed version of thisdesign has garnered its share of”wives’ tales” over the 30-plus yearsof commercial manufacture. Follow-ing are the ”Top 10” notions and realissues that generate continuous dis-cussion among users of these prod-ucts. Although many of the topics areworthy of individual papers, we arelimited here to an overview alongwith some helpful hints.

10. EFFICIENCY: HERE TODAY, GONE TOMORROW?

Partial emission pumps use openimpellers and therefore do not rely

on wear ring clearances to maintainhydraulic efficiencies. Although thisdesign allows more recirculationback to the suction than an enclosedimpeller design, efficiency definitelydepends on the disk friction compo-nent. Disk friction is the drag lossbetween the body of rotating fluidbeing carried by the impeller and thestationary walls of the chamber(Figure 2). In clean, non-corrosive

streams, the finish on the chamberwalls remains in ”as new” conditionresulting in steady, long termhydraulic efficiency.

Typical finishes are machined toa 62 (micro inch) RMS value. Thus, asurface finish of 250 can reduce thehydraulic efficiency by as much as 10points. In extreme cases, total powerconsumption has been known to dou-ble that of the ”clean” rating. Of the

High Speed/Low FlowPumps: Top 10 Issues

The authors answer your top ten questions about high speed/low flow pumps that perhaps you were afraid to ask.

BY: DAVE CARR AND ROBERT LINDEN

T

PHOT

O CO

URTE

SY O

F KO

P-FL

EX, I

NC.

Photo 1. A centrifugal pump performing high speed/low flow dutiesin a Gulf Coast plant.

Page 125: Centrifugal Pumps Handbook

126 The Pump Handbook Series

two chamber surfaces, the finish ofthe backing plate is actually of greatersignificance than the impeller bowl inits effect on efficiency. Machinedskin cuts at depths of 0.005 to 0.010inch have been shown to efficientlyrestore the pump’s original perfor-mance by reconditioning the surfacesto the material’s original finish.

In addition to wear of the bowlmaterial, there is also potential forforeign material build up on the sur-faces. The corresponding skin, orfilm, similarly changes the effectivefinish and may also degrade the effi-ciency. The pump’s sensitivity to thiscondition increases with tip speedand operation with lighter liquids.For this reason, wetted materials maybe selected to shed any build up or toresist the potential for accumulationsto occur. For example, salts have anaffinity for carbon steel materials.Build-up of the type commonly expe-

rienced in contaminated butanestreams has been negated by upgrad-ing to 316 stainless steel construction.

Either of these effects can be ver-ified by measuring the motor amper-age or the liquid’s temperature rise.The latter indicator increases propor-tionately to the decrease in hydraulicefficiency, i.e. by the equation:

∆T={H(1-η)/778Cpη}

Units: T is in °F, H is head infeet, Cp is specific heat in BTU/LB-°Fand η is efficiency expressed as a dec-imal.

Comment: Practically speaking,temperature rise analysis can be diffi-cult because it typically rangesbetween 2 and 10°F.

9. THE SUCTION SPECIFIC SPEED IS WHAT?

Quoting from Lobanoff & Ross,”The inducer is basically a high specif-ic speed, axial flow pumpingdevice...that is series mounted preced-ing a radial stage to provide overallsystem suction advantage.” This rela-tionship is paramount to evaluatingsuction specific speed with high-speedpumps. Figure 3 shows a portion of awell-rounded inducer family.

An industry rule of thumb is tolimit suction specific speed to a valueof 11,000 (English units) in heavyduty process pumps. It is based onthe premise that pumps operating atgreater suction specific speeds have

demonstrated a shorter service life,most likely due to rougher off designoperation.. As conventional impellersare capable of suction specific speedvalues greater than 11,000, via over-sized impeller eye configurations, thismay be a good guide. To provide auseful net positive suction head(NPSH) performance guide for high-speed pumps, however, will require achange in the reference point!

High-speed pumps inherentlyrequire inducers to achieve competi-tive performance. Inducer designshave advanced to the point of provid-ing reliable, cavitation-free operationwith suction specific values ofapproximately 24,000. Efforts todesensitize inducer operating rangeshave been addressed by optimizingblade number, angle and passageareas as well as using various inletbypass designs. One design thatresists cavitation surge, referred to asa ”backflow recirculator,” is reportedto improve inducer turndown to anear shut off condition. Suction vor-tex breakers are also used but are lesseffective than flow stabilizers.Nonetheless, they can improve lowflow stability, with high-speed/induc-er style pumps, by as much as 25 to35%.

In general, inducer use is dis-couraged with conventional APIpumps, but high-speed pumps clearlyrequire them. Several leaders in theHydrocarbon Processing Industry

FIGURE 1. OPEN IMPELLER

Back CoverSurface

Impeller BowlSurface

FIGURE 2. PARTIAL EMISSION CHAMBER SURFACES

Inducer CInducer B

Inducer A

GPM / RPM

Suct

ion

Spec

ific

Spee

d

FIGURE 3

Page 126: Centrifugal Pumps Handbook

The Pump Handbook Series 127

(HPI) have recognized the uniqueposition that this equipment occupiesin the marketplace and have explicit-ly exempted such designs from the11,000 suction specific speed limit.Care still must be exercised, howev-er, in matching particular inducerconfigurations with the pump’s oper-ating flow range, and interaction withmanufacturers that recognize cavita-tion erosion limits within their appli-cation guidelines.

8. HOW CAN MECHANICAL SEALSHANDLE THAT SPEED?

Mechanical sealing of high-speedpumps presents numerous obstaclesincluding the potential for vibration,high sliding speeds, heat generationand high sealing pressures. Much ofthis is the direct result of rotationalspeeds that reach a maximum of25,000 rpm, which is more commonwith compressors than with pumps.These conditions place high-speed

pump manufacturers in the uniqueposition to accept ”seal success own-ership.” The bulk of the issues areaddressed by reversing the conven-tional seal configuration. Vibrationcontrol dictates the spring-loadedcomponent as the stationary partwith the hard face as the rotatingmember. Small seals, typically witheither 1-1/4” or 1-1/2” diameters, areused to minimize sliding speeds. A 1-1/4” seal at 15,000 rpm has a facesliding speed of approximately 82 feetper second. This equates to a 5-1/2”seal operating at a traditional 3,550rpm (Figure 4.). When combinedwith the fact that open wheelimpellers present seal pressures near-er to suction than to discharge (i.e.,approximately 10% rise over the suc-tion pressure) manageable pV valuesare seen at the seal faces. This allowsmost applications to be handled byconventional carbon vs. tungsten car-bide material combinations. Silicon

carbide also may be used to raise theoperational limits.

Most high-speed pump seal prob-lems are not caused by high speed,but by a lack of understanding orinformation regarding the fluid’sproperties. The most common prob-lems are rust and scale, inadequatevapor pressure margin and a build upof solute at the atmospheric interface.Proper application of typical API sealflush plans, e.g., -31 (flush through aseparator), -13 (reverse flush) and -54(quench), normally negate these con-cerns and promote good seal reliabili-ty with high speed pumps.

7. PREPARING THE PUMP FORSTARTUP.

High-speed pumps are typicallyaccorded extra care during the start-up process due to respect for thetechnology. The lubricant level, oilcooler venting, fresh oil filter installa-tion (if applicable) and driver rotationare inspection points which are con-sistently addressed. Seal piping, how-ever, can be quite another story. APIStandard 610 dictates, and industrypractice provides, that all seal portsbe plugged prior to factory shipment.In the field, the permanent appear-ance of some of these plugs frequent-ly leads to failure to remove them.The pump case tags, engineeringdrawings and the instruction manualall provide a critical definition of theports’ functions and the correspond-ing auxiliary piping requirements,and must be followed.

Unfortunately, failure to proper-ly configure the pump often will notcause any immediate problems. Acommon occurrence is an improperlyvented port that can direct accumu-lated process seal leakage toward theback side of the gearbox mechanicalseal, or bearing seal. The potential forlubricant contamination exists whenan atmospheric drain is connected toa flare header that experiences signifi-cant upsets. Care should be exercisedto ensure that these actions do notoccur since the livelihood of the gear-box depends upon limiting this pres-sure to an absolute maximum of 10psig.

6. SUBTLE TRUTHS ABOUT NPSH.Net positive suction head (NPSH)

is a subject worthy of its own full

1. 1 1/4", 1 1/2", & 2" SIC. vs. C.

2. 1 1/4 IN. T.C. vs. C.

3. 1 1/2 IN. T.C. vs. C.

4. 1 1/4 IN. T.C. vs. SIC.

5. 1 1/2 IN. T.C. vs. SIC.

6. 2 IN. T.C. vs. C.

7. 2 IN. T.C. vs. SIC.

SIC. = SILICON CARBIDET.C. = TUNGSTEN CARBIDEC. = CARBON

25,000

20,000

15,000

10,000

5,000

00 500 1,000 1,500

SEAL DIFFERENTIAL PRESSURE (PSIG)

PUM

P SP

EED

(RPM

)

FIGURE 4. PRESSURE-SPEED LIMITS AS A FUNCTION OF SEAL FACE MATERIALS

1.

2.

3.4.5.6.

7.

Page 127: Centrifugal Pumps Handbook

128 The Pump Handbook Series

fledged article, and has been coveredpreviously in Pumps and Systems. Itshould be emphasized, however, thata centrifugal pump’s NPSH perfor-mance is established based upon thebreakdown of the standard head ver-sus capacity curve. When inducersare used, the total pump NPSH per-formance is measured, not just theinducer’s performance. The mostcommonly accepted parameters arebased on the Hydraulic Institute’s 3%head suppression criterion. By defini-tion, however, the pump will cavitatewhen conditions at the suction flangemeet those test, or predicted curve,conditions.

High-speed inducers (in essenceaxial pumps) can develop as much as25 to 100 feet of head. This energy,however, cannot be included withinthe overall pump requirements dueto corresponding inlet eye losses atthe centrifugal impeller. The netresult is that the centrifugal portion ofthe pump still must be sized for thefull design head rating.

The subtle aspect of NPSHrevolves around the pumped fluid’sproperties and its potential to flash.Cavitation damage is a function of aliquid’s propensity to release vapor.Water pumps typically will producerated head and flow, albeit with thepotential for some material damage,despite close proximity to the 3% headsuppression value. This may be attrib-uted to the high surface tension charac-teristic of water. The same trait makesit aggressive toward cavitation damagebut the pump generally works!

Difficulty occurs at the other endof the spectrum where high speedpumps are used with low specificgravity services. Using the 3% sup-pression value for light fluids, plusthe industry rule of thumb for anadditional 2 to 3 feet safety factor, theNPSH margin may not be adequatewith low specific gravity fluids. Oneexample of this situation is realizedwhen a slight heating of the fluidoccurs on the suction side of thepump, particularly with above-ground supply pipes from storagevessels to transfer pumps. This canresult in off-gassing and failure of thepump to hold prime. Ironically, thissituation contradicts the API hydro-carbon offset factors that are typically

prohibited by company specifica-tions. In general, increasing theNPSH margin by an additional 4 to 6feet is appropriate with light gravityliquids, i.e., less than 0.7 specificgravity.

Also of concern is the fact thatthe NPSHR value typically increasesbeyond the best efficiency point ofthe machine. Pump startups are com-monly uncontrolled and result inoperation at the end of the curve dueto a lack of sufficient back pressure.The practical result of this situation isthe fact that the pump may then runat too high an NPSH requirement,and could promote a disconnectbetween the impeller/inducer and theliquid stream. Therefore, the controlvalve’s initial trim position and man-ual venting of the pump systemshould be anticipated in preparationfor startup.

5. CENTRIFUGAL PUMPS IN APOSITIVE DISPLACEMENT WORLD.

High-speed centrifugal pumpsare often installed in applicationsdesigned for positive displacement(PD) pumps. This is due to theinherent ability of both pumps todeliver a high differential pressure.Unfortunately, the two designs mustoperate under significantly differentcontrol schemes. This fact must berecognized when retrofitting fromone configuration to another.

Figure 5 shows the theoreticalcharacteristics of PDs and centrifu-gals with regard to flow and headcapabilities. It is evident that the posi-tive displacement design is limited byhead (pressure) and the centrifugal byflow. Consequently, the PD uses apressure relief valve to prevent overpressurization and to bleed off excesscapacity. In practice, centrifugalpumps exhibit only a moderate headrise across their operating region. Theradial vaned centrifugal, in particular,demonstrates a 5-10% head rise fromthe best efficiency point (B.E.P.) tothe peak of the head versus capacitycurve. This margin does not facilitatethe use of a pressure relief system forcontrol purposes. Further, the reliefvalve scheme can result in wastedpower when the pump is allowed torun out to the extreme right of theB.E.P., under ”low load” conditions.

The recommended high-speed pumpcontrol system is with the use of flowcontrol. At first glance, this conceptcan be intimidating but is essentiallysynonymous with level and masscontrol schemes that are typical with-in process systems. Some processesdo demand strict pressure control.When that is the case, a pressure con-trolled throttle or bypass source maybe required.

4. WHAT ABOUT UNCONTROLLEDFLOW OPERATION?

The effect of low flow operationon centrifugal pumps is commonlydiscussed, but rarely is the oppositeend of the performance curve consid-ered. Regardless of the hydraulichardware, NPSHR generally increas-es with a pump’s operation at greaterthan its rated flow. This topic hasbeen mentioned within issue #6 andtherefore will not be discussed fur-ther. Single volute and diffuser stylepumps experience increased radialloads when applied at greater than

Flow Coefficient

FIGURE 5. THEORETICAL HEAD RISE

Radial Vaned Centrifugal

Positive Displacement

Backward Lean Centrifugal

Head

Coe

ffici

ent

Load

FIGURE 6. CENTRIFUGAL PUMP RADIAL LOAD CHARACTERISTICS

Percent of Design Flow

Single Volute

Diffuser

Desi

gn

0 25 50 75 100 125 150

Page 128: Centrifugal Pumps Handbook

The Pump Handbook Series 129

design flow rate (Figure 6). It is gener-ally understood that sufficient bear-ing over capacity, and/or controllimits, must be applied to the pumpto account for these events. High-speed centrifugal pumps share thesesame basic design needs, but alsomust address a phenomenon oftenreferred to as discharge cavitation.

Discharge cavitation occurs with-in single divergent conical (pointemission) diffusers when the pump isoperated at rates to the right of kneeof the performance curve (Figure 7).Under such conditions, a low pres-sure zone forms on the trailing edgeof the impeller’s blades. Vapor bub-bles are formed and subsequentlycollapse in a manner consistent withthe standard definition of cavitation.The result is impeller blade pitting,increased pump vibration and theclassic ”rock pumping” noise associat-

ed with suction performance prob-lems.

Such an occurrence is commonwith systems where the pump hasbeen oversized on head, inadequatecontrol is afforded to maintain opera-tion within a maximum flow limit, orsimply during a process startupwhere the system is being filled. Agood rule of thumb is that this type ofpump should be controlled to a maxi-mum flow of 120% of the pump’s rat-ing. System requirements exceedingthis value require a conversion to alarger diffuser throat to accommodatethe actual process demand. Use ofthe design’s conversion capabilities isalways preferable to grossly oversiz-ing the machine for future require-ments.

3. A CONFLICT BETWEEN CURVES!Some, but not all, high-speed

pumps produce flat or ”drooping”curves. This characteristic, which iscommon to low specific speed pumps,has elicited much discussion regard-ing their inherent ability to be con-trolled. Such pumps, however, havebeen successfully applied in tens ofthousands of applications in spite ofthese concerns. Unfortunately, how-ever, many times this has beenaccomplished at the expense of ener-gy by oversizing the pump’s ratedhead and then employing a dischargeorifice to artificially steepen the curveas it runs back toward its minimumflow point.

The application key is to under-stand the pump and the systemcurves. It is a common belief thatdrooping curves are difficult to con-trol because the pump has two flowpoints associated with a single head(pressure) point. The result is a ten-dency for the pump to ”hunt”between the two flows. What is oftenoverlooked, however, is that thepump merely reacts to what the sys-tem presents it with (i.e., it operatesat the exact point where the systemand pump curves intersect).

A process system’s characteristicresistance curve typically is made upof two components. The first isreferred to as the ”fixed” element andis associated with the system’s staticcomponent, e.g., the operating pres-sure of a process tower. This element

is considered to be constant withrespect to the flow rate of the system.Conversely, the ”variable” compo-nent may be simply thought of as thefrictional element that is related tothe pumped flow rate. These two fac-tors are shown in Figure 8. Whencombined, they become the basic sys-tem curve.

Figure 9 shows a typical head ver-sus capacity curve (ABC) with adrooping characteristic. Point B signi-fies the best efficiency point (B.E.P.),point A the cutoff flow and point C isthe stonewall condition. Superimposedon this curve is the basic systemcurve (SB) which was derived fromthe previous discussion. Without sup-plemental pump control, the systemwill demand a flow rate equal to XB.The system head curve can be modi-fied with changes to the piping sys-tem or by regulating the pressuredrop across a control valve. The latterapproach is a typical means of con-trolling centrifugal pumps and yieldsnew system curves as indicated by(SD) and (SA). It is seen that thepump’s head capability is equal at theXA and XB flow points yet successfulpump operation is accomplished as aresult of the modified system curve.

We would be remiss in not point-ing out that this type of curve doeshave its shortcomings. First, pressurecontrol normally is not practical dueto the relatively small head rise thatoccurs between the rated and maxi-mum head points, typically on theorder of 5-10%. This fact favors theuse of flow control methods. Second,systems that are comprised predomi-nantly of the ”fixed” component, i.e.,those that exhibit little influence as aresult of the system’s demand flow,

Resi

stan

ce

Flow

B

V

F

S

S'

B = System CurveV = "Variable" ComponentF = "Fixed" Component

FIGURE 8

Head

50 100 120Percent of Design Flow

FIGURE 7. POINT EMISSION PUMPPERFORMANCE CHARACTERISTICS

Knee of Curve

Potential Zone ofDischarge Cavitation

S

A BD

CXA XD XB

FIGURE 9

Page 129: Centrifugal Pumps Handbook

130 The Pump Handbook Series

may result in an undersized pump ifan active appraisal of the pump/sys-tem interaction is not performed.

Regardless of the centrifugalpump type, a discrete intersectionbetween the pump and systemcurves will always complementpump stability and controllability.Conscientious attention to the interac-tion between pumps and systems cantame both so that they work in har-mony.

2. WHY WON’T THIS PUMP WORK ONWATER?

Pump manufacturers typicallyuse water as a performance testmedium for safety and conveniencereasons. Many pumps, however, aresold for process liquids that varybetween a 0.4 and 0.8 specific gravi-ty. Factory tests often use one-halfspeed motors, or other speed changesto compensate for the increasedpower that results from water’s den-sity difference to the contract liquid.Field operation, however, commonlyuses an initial run-in on water toflush and prove the system. Two typ-ical repercussions of this action arean expected overload of the driver oran unexpected overload of a high-speed gear or bearings. The high-speed pump is particularly vulnerableto this off-design operation as a resultof its common use in light hydrocar-bon processing applications and as aresult of small, custom matched,components which capitalize on thespecific application’s needs.

Throttling the pump may or maynot meet the pump’s needs becausebearing loading may be violated inextreme cases. The moral of the storyis to check with the pump vendorbefore proceeding with off-spec tests.The optimal approach is to advisehim of your complete run-in condi-tions prior to placing a pump order toensure that it will meet all of theintended uses.

1. WHERE IS THAT CONTROL VALVE?Centrifugal pump designers

expect throttle valves to be very nearthe pump discharge, while systemdesigners prefer a location near thedemand point. This issue becomesmore than one of aesthetics when itinvolves high energy pumps.

A surge phenomenon may occurwith pumps with either continuouslyrising or drooping head versus flow

curve shape attributes. It is distin-guished by fluctuations in headcapacity at the pump’s low flow con-ditions. The combination of high-speed technology with relatively lowdesign flows introduces unique chal-lenges to the pump designer anduser.

Empirical testing shows that thispump design’s low flow stability isdirectly influenced by the systemwithin which it operates. The singlepoint emission diffuser and a Barskeimpeller may be simplistically charac-terized as discharging flow each timea blade passes by the throat. It is rea-sonable to envision a void betweenthe time in which one blade passesand the next arrives to distribute itssupply. This interim period representsthe opportunity for the discharge con-trol valve’s location to influence thepump’s low flow stability.

The valve’s interactive processmay be visualized in the context of asimple spring mass system. The liquidin the system acts as the mass and allstorage devices within the system,e.g., piping, vessels, etc., provide thespring medium. Excitation of this sys-tem may be initiated from a numberof sources but often may be related tothe blade passing frequency. Thegreater the energy that is accumulatedwithin this system, i.e., the spring, thegreater the propensity for it to disruptthe pump’s stability. The momentaryflow reversals cause surge circulationbetween the blades. This type of sys-tem is sometimes referred to as ”soft”or ”spongy” since it reinforces theamplitude of the theoretical springforce. The destructive energy of thissituation is greatest with an increasingmass of liquid, when the throttlevalve is remotely located, and withincreasing pump power.

This phenomenon is minimizedwhen the system can be described as”hard.” This is accomplished by plac-ing a control valve, or orifice, nearthe pump’s discharge flange. Ineffect, this change reduces the liquidmass, thereby minimizing the ampli-tude of the spring’s movement, andthe flow oscillations. The valve/orificealso disrupts the frequency of theexcitation force and further improvesthe pump’s low flow stability.

The ”hard” system should alwaysbe the goal since it discourages theformation of a potentially dangerousenergy source that can damage piping

and induce mechanical vibration intothe high-speed pump. Good rules ofthumb are that transmitted powerlevels of less than 25 horsepower areminimally affected by this phenome-non and the optimum control valvelocation is within 5 feet of the pump’sdischarge flange. Failure to addressthis situation can reduce a 200 horse-power pump’s minimum recom-mended continuous flow rate from 40to 65% of the B.E.P., based upon thevalve’s placement 25 feet, rather than5 feet, from the pump’s dischargeflange. ■

REFERENCESVal S. Lobanoff and Robert R.

Ross, Centrifugal Pumps Design &Application, Gulf Publishing Company,Houston, TX, 1985.

Donald P. Sloteman, Paul Cooperand Jules L. Dussord, Control ofBackflow at the Inlets of CentrifugalPumps and Inducers, presented at theFirst International Pump UsersSymposium (1984).

Robert Linden is the director ofSundyne and Sunflo products forSundstrand Fluid Handling, Arvada,CO.

Dave Carr is a senior marketingspecialist with Sundstrand FluidHandling.

Page 130: Centrifugal Pumps Handbook

The Pump Handbook Series 131

CENTRIFUGAL PUMPS

HANDBOOK

All pumps are pressure rated.The rating is the maximumpressure a pump casing cansafely contain, often termed

maximum allowable working pres-sure or MAWP, at a given tempera-ture (Figure 1). Temperature affectsthe rating because the strength andstiffness of the materials used for cas-ings – mostly metals – vary with tem-perature. From that simple definition,the subject becomes more complicat-ed as we add definitions for thepump’s maximum discharge pres-sure, the means of correcting for dif-ferent temperatures, and the questionof what to do with the casings ofpumps that develop high differentialpressures.

The maximum allowable work-ing pressure (MAWP) of a pump’scasing is a function of its geometry,the material from which it is fabricat-ed and the intended service tempera-ture. Before delving into numbers,let’s address a fundamental aspect ofcasing geometry, namely the casingjoint.

THE CASE FOR CASING JOINTS

Pump casings must have a joint— either at right angles to the shaft

axis (radially split) or parallel to it(axially split) — to allow the pump tobe assembled and dismantled. Mostpump casings are radially split: insmall overhung pumps, either hori-zontal or vertical, because of theirinherently lower cost; in mediumsize and large vertical pumps for bothlower cost and ease of maintenance;and in high pressure pumps, eitherhorizontal (Photo 1) or vertical,because it’s the more cost effectivesolution. In between these extremes,axially split casings are preferred forhorizontal single stage double suctionand multistage pumps (Photo 2) forlower cost and ease of maintenance.They are, however, limited in thepressure rating they can economical-ly achieve.

API-610, 7th Edition1, recom-mends using radially split casings forhydrocarbon service when the maxi-mum discharge pressure is above1,000 psig (70 bar), the pumpingtemperature above 450ºF (235ºC), orthe liquid specific gravity below 0.7.These conservative limits reflect thedifficulty various refiners have hadmaintaining pressure tightness inaxially split casings. There are, how-ever, many examples of axially split

casings being used successfully inhydrocarbon service beyond theselimits, and the forthcoming 8thEdition of API-610 will recognize thisby raising the recommended pres-sure limit to 1,450 psig (100 bar). Inwater injection and boiler feed appli-cations, axially split casings are regu-larly used to pressures of 2,500-2,750psig (175-190 bar). Higher pressuresare possible, but the cost of the cas-ing can become prohibitive, andmaintenance of the split joint gasketa major concern.

The radially split casing inPhoto 1 is one piece with a cover orhead at the outboard end. Thisdesign has one high pressure seal,and the pump can be dismantledwithout breaking its suction or dis-charge connections or moving its dri-ver. There is an alternative form ofradially split casing, known various-ly as “ring section” or “segmentalring,” composed of many piecesclamped together with tie bolts(Photo 3). This design achieves lowcost at the expense of maintenance.It has many casing joints, does notcomply with API-610 and was wise-ly dropped from use in the US in themid 1930s for all but small industrial

Pump Ratings VitalWhen Pressure’s OnThough a relatively simple subject, pump ratings can generate much disagreement.Using objective evaluation procedures, however, we can shed light on the topicwithout generating heat and pressure.

By J. T. (“Terry”) McGuire

Page 131: Centrifugal Pumps Handbook

132 The Pump Handbook Series

boiler feed pumps.

DETERMINING MAWP

For a given casing geometry, apump’s MAWP is determined byallowable stress unless strain (deflec-tion) at critical sealing surfaces dic-tates a lower stress. The allowablestress may be that from ASMESection VIII, Division 12, as requiredby API-610, or some other similarlimit. Designs using the ASME stresslimits also include a casting integrityfactor, which is 0.8 unless volumet-ric NDE of the castings allows ahigher factor. Note that the allow-able stress from ASME Section VIII,Division 1, includes a large design or“ignorance” factor to account fordesign using simple means of esti-mating the stress. When moresophisticated means of estimatingthe stress are employed, such asfinite element analysis (FEA), using ahigher allowable stress is justifiedbecause the local stress values are

Photo 1: High pressure radially split casing

Photo 2: Axially split multistagepump casing

Photo 3: Segmental ring casing

Page 132: Centrifugal Pumps Handbook

The Pump Handbook Series 133

now known with quite high accura-cy. In such cases, the allowablestress from ASME Section VIII,Division 2, can be used providedthat the FEA model has been veri-fied for accuracy and that appropri-ate material quality is used.

CALCULATING MAXIMUM ALLOWABLEDISCHARGE PRESSURE

In operation, the maximum dis-charge pressure developed by a cen-trifugal pump is equal to the sum ofthe maximum suction pressure itcan be exposed to plus the maxi-mum differential pressure it candevelop. In a simple world, that defi-nition would be sufficient, but theworld’s not so simple. Pumps arepurchased with various reserves,margins and tolerances in head androtative speed that complicate thedefinition of maximum dischargepressure. Following the require-ments of API-610, 7th Edition, pro-duces two definitions, one for fixedspeed pumps, the other for variablespeed. With Figure 2 as a reference,the definitions are:

Fixed speed pumpPd, max = Ps, max + 1.05(∆ Pmax)fH (1)

Variable speed pumpPd, max = Ps, max + ( ∆Pmax)fH(fN)2 (2)

Where:Pd, max is the maximum discharge pressure

Ps, max is the maximum suction pressure

∆ Pmax is the maximum differential pres-sure at rated specific gravity

fH is a factor to account for the allowabletolerance on shut-off head

fN is a factor to account for the allowableoverspeed to trip.

API-610’s requirement for vari-able speed pumps covers the headreserve in the allowable positive head

tolerance, and it assumes the pumpwill be operated as if fixed speed, i.e.its flow will be controlled by throt-tling. Boiler feed pumps for centralstations are often variable speed driveto avoid the losses associated withcontrol by throttling. As such, theirhead always corresponds to the sys-tem resistance at any given capacity,and it’s therefore appropriate to use adifferent definition for the maximumdischarge pressure. Referring toFigure 3, it is:Pd, max = Ps, max + PCMR (3)

Where: PCMR is the differential pressure

at the pump’s continuous maximumrating (CMR).

With this definition, the 10-15%higher discharge pressure that couldbe developed in the event the driverwent to overspeed while the pumpwas blocked in is classified as amomentary excursion into the marginprovided by the casing’s hydrotestpressure.

Positive displacement pumps,unlike centrifugal pumps, will devel-op pressure equal to the resistancethey encounter, up to the mechani-cal limit of the pump or its drive.This is obviously an extremely dan-gerous possibility, and so the cardi-nal rule in the application of positivedisplacement pumps is the provisionof a full capacity relief valve at theirdischarge, upstream of any possibleobstruction. The accumulation pres-sure of the relief valve, the addition-al pressure drop across the valve todischarge its rated capacity, shouldbe no more than say 20%. With thisprovision, the pump’s maximum dis-charge pressure is (Figure 4):

Pd, max = PsetfA (4)

Where:Pset is the relief valve set pressure

fA is a factor to account for the accumula-tion pressure.

PUMP RATING VERSUS REQUIRED

For a pump’s pressure rating tobe acceptable, the MAWP of its cas-ing, at the intended service tempera-ture (Figure 1), must be at least equalto the pump’s maximum dischargepressure, as calculated by the applica-ble equation above. Depending uponthe design status of the pump beingconsidered, there are two approachesto achieving this. For existing designs,which is the usual case, the MAWP ofthe pump’s casing at the design tem-perature, generally 100ºF (40ºC), iscorrected to that for the intended ser-vice temperature, T, by the equation:

MWAPT = MWAPD(σT/ σD) (5)

Where:σT is the allowable stress at the intended

service temperature

σD is the allowable stress at the designtemperature Unless specifically stated other-

wise in the engineering specification,the intended service temperature forcasings is taken as the pump’s normaloperating temperature. The rationalefor this is that the maximum temper-ature generally represents a possibleshort term transient condition.

When the casing is beingdesigned specifically for the applica-tion, a common practice with fabri-cated casings, the pump’s maximumdischarge pressure is used to calcu-late a minimum design pressure,MDP, by the equation:

MDP = Pd, max(σD/ σT) (6)

The MDP is generally roundedup to the nearest common increment,25 psi being the ASME practice, or toa minimum pressure required by theengineering specification or a con-nected flange.

MINIMUM CASING DESIGN PRESSURE

Page 133: Centrifugal Pumps Handbook

134 The Pump Handbook Series

FOR SHOP TESTING

When designing a pump to han-dle a liquid of low specific gravity, itmay be necessary to raise the MDPcalculated from equation (6) to accom-modate the pressure that will bedeveloped during shop testing withcold water. The equation for the max-imum discharge pressure on test is:

Pd, test = Ps, test + (Hmax/2.31)fH (7)

Where:Ps, test is the maximum suction pressure

during the testHmax is the highest head the pump will

develop during the test, in ft.

The maximum pump headdeveloped during the shop test isusually at shutoff, except in the caseof a pump with a drooping headcharacteristic, or high energy multi-stage pumps, which are tested onlydown to their minimum continuousflow. The head tolerance factor, fH, isthat for shutoff.

CONTROVERSY OVER DUAL PRESSURECASINGS

Multistage centrifugal pumpsdevelop large pressure differentials.This means various regions of their

casing normally are subjected to dis-tinctly different pressures. Once thepump exceeds a certain size and pres-sure rating (3-inch discharge andANSI 1,500 # flanges are a good start-ing point) it becomes more economi-cal to design the casing for twopressures, a convenient low pressurefor the regions subjected to suctionpressure, maximum discharge pres-sure for the remainder of the casing(Figure 5). This practice is normal inthe utility industry and is recognizedby API-610, the standard for theprocess industry. API-610 also recog-nizes this is a controversial topicwithin the process industry, so itincludes the option of specifying thatthe entire casing be designed for max-imum discharge pressure.

Designing the normal low pres-sure regions of a barrel pump casingfor discharge pressure has two nega-tive effects that need to be taken intoaccount. The first is that the shaft sealdesign is often compromised. Hotcharge pumps, for example, are bestequipped with metal bellows typeshaft seals. These cannot withstandstatic pressures of more than 350-400psig (24.0-27.5 bar). Therefore, push-er type seals have to be used. Pusherseals rely on some form of cooling topreserve their elastomer dynamicgasket so the design has becomemore complicated and potentially lessreliable. The second drawback is thatthe flange of cartridge mounted sealsbecomes so large and heavy thatinstalling and removing it presents aserious handling problem.

The controversy over dual pres-sure casings appears to have its originin operating practices and the staticpressure tightness of mechanicalseals. In the utility industry andabout half of the process industry,standard isolation practice is to blockin the discharge, then open a drainbefore blocking in the suction. Whenthis sequence is followed, the entirecasing and suction piping back to the

FIGURE 2

Allowable actual pump head curve, API-610, 7th Edition

Flow

Tota

l Hea

d Proposal curve

Ratedpoint

Allowable test curve with maximumpositive tolerances

FIGURE 3

Tota

l Hea

d

Flow

Variable speed boiler feed pump

Bypass

H @ minimum speed

H @ rated speed

Systemhead

CMR

FIGURE 1

Variation of MAWP with pumping temperature

Pumping Temperature

Pres

sure

MAWP

Flow

FIGURE 4

Pressures in positive displacement pump applications

Pres

sure

SlipRelief valve accumulation

Relief valve setpressure

System resistance

Suction

Page 134: Centrifugal Pumps Handbook

The Pump Handbook Series 135

suction block valve cannot be acci-dentally subjected to discharge pres-sure by a small leak past thedischarge valve. The alternativesequence, blocking in the suctionbefore opening a drain, when carriedout on pumps with mechanical seals,does risk subjecting the entire pumpcasing and suction piping down-stream of the suction block valve todischarge pressure. There have beeninstances where doing this has rup-tured the suction piping or the pumpcasing. It does seem that a judiciouslyplaced burst disc or relief valve couldpractically eliminate the risk of mis-adventure without compromising thedesign of the shaft seals.

HYDROSTATIC TEST PRESSURE

A casing, or the various regionsof a dual pressure casing, is hydrosta-tically tested at

Phydro = 1.5 Pdesign (8)

Pdesign is either MAWPD for pre-engi-neered casings (see Equation 5), orthe greater of MDP from Equations 6or 7 for engineered casings.

REVIEWING THE OPTIONS

Casings with the highest pres-sure ratings are radially split. Thereare actual examples of these designedfor 14,000 psig. Axially split casingsare preferred for large axis singlestage double suction pumps and hori-zontal multistage pumps because ofease of maintenance. They cannotalways be used, however, becausetheir pressure rating is currently lim-ited to 1,000 psig by API-610, 7thEdition, for hydrocarbon service.

This is due to be increased to 1,450psig effective with the 8th Edition ofAPI-610, although there are examplesof their operation at 2,500 to 2,700psig for boiler feed and water injec-tion services.

Higher pressures are possible,but the cost of the casing can becomeprohibitive, and the maintenance ofthe split joint gasket a major concern.Single stage, single suction pumps aregenerally radially split – for econom-ics in the smaller sizes, both horizon-tal and vertical, and for ease ofmaintenance in the very large verticalaxis designs.

When a methodical approach istaken, the issue of pump pressure rat-ings is not too difficult. This evalua-tion and selection process can befurther simplified by following theevaluation procedure outlined in thebox accompanying this article. ■

REFERENCES:

[1] API-610, 7th Edition,Standard for Centrifugal Pumps forGeneral Refinery Service, AmericanPetroleum Institute, Washington, DC,1989.

[2] ASME Boiler and PressureVessel Code, Section VIII, Divisions 1

A = suction pressure + 75 psiB = maximum discharge pressure

Regions of barrel pump casing subjected to different pressures.

A AB

FIGURE 5

Evaluation ProcedureAnswering the following questions in sequence will help avoid errors:Is the preferred form of casing joint suitable for the intended pres-

sure, temperature and liquid specific gravity?Is the maximum discharge pressure, calculated as required by the

applicable industry standard, less than the casing’s MAWP at the intendedservice temperature?

Is there a minimum industry or code pressure rating for this class ofpump?

Do the casing flanges or nozzles have a pressure rating at leastequal to that of the casing region to which they are connected?

Is the minimum design pressure of the casing determined by the pres-sure developed during shop testing with water?

Does the purchaser’s specification require all regions of multistagepump casings to have a MAWP equal to or greater than the maximum dis-charge pressure?

Page 135: Centrifugal Pumps Handbook

Many industry pumpproblems are notcaused by im-proper operationor faulty mainte-nance, althoughthese are the pri-mary focus of most

reliability improvement efforts. Inreality, most problems are traceable toimproper initial application or changedoperating conditions. The pump is justnot right for the job. Or it once was,but is no more. Misapplication can beavoided, however, by observing threeimportant principles: communication,communication and communication.

In the past, and sometimes eventoday, pump users assume an adver-sarial relationship with their suppli-ers. Mutual suspicion is the order ofthe day, with each side trying to“win” or gain an advantage over theother. On the other hand, wise pumpusers are treating suppliers as valu-able resources. They seek win-winoutcomes where everyone benefits.Pump manufacturers, too, havefound that making today’s sale is notas important as building a long termrelationship with customers. Out ofthis industry-wide shift in attitude hascome a new opportunity for suppliersand users to work together to put thebest pump in a given application.

With all of today’s corporatedownsizing and re-engineering,pump users cannot afford to ignorethe wealth of help available frompump suppliers. After all, supplierscan draw from a broad spectrum ofindustry experience to help a usersolve a particular problem. Thisinformation exchange benefits usersby taking solutions developed in oneindustry segment and introducing it

to other sectors. Pump suppliers ben-efit by getting valuable feedback ontheir designs so they can improvetheir products and broaden theirapplications. And as this articleexplains, communication is the keyto making it happen.

There are two primary areas ofpump user/supplier interaction. Thefirst is when the user initially pur-chases a new pump. The second iswhen a user attempts to improve theperformance of an existing pump.These two activities are different yethave much in common. The mostimportant of these is the need forclear, open and honest communica-tion between users and suppliers.

NEW PUMP PURCHASESPurchasing a new pump

requires a team effort between cus-tomer and supplier. Much effort hasbeen expended over the years todevelop specifications and industrystandards such as API and ANSI.These standards form the founda-tion of a new pump purchase. Fol-lowing are some additional ideasthat will help you select and pur-chase the best pump for a givenapplication.

PROCESS DATAAccurate process data is needed

to achieve a successful pump applica-tion. Without good information, thepump supplier is already fighting withone hand tied behind his back, andthe battle hasn’t even started. A wholebook could be written on the subjectof properly sizing pumps, but the cal-culation and engineering aspects arebeyond the scope of this article.

Good communication is the keyto getting accurate process data onto

the pump data sheet. Typically, theflow rate is established by theprocess for which the pump is beingselected. A calculation then deter-mines the discharge pressurerequired to move this amount of flowthrough the piping system. Thisprocess is fairly straightforward, butpitfalls exist even at this elementarystage. Many process design engineersdo not understand that a typical cen-trifugal pump has an operating rangeof only 40–120% of Best EfficiencyPoint (BEP) flow rate. This may leadthem to set the rated flow rate of thepump higher than required, possiblyto allow for future expansion of theunit. Unfortunately, this futureexpansion either is many years awayor never happens. As a consequence,the process unit is left with a pumpthat is operating at, or at much lessthan, the lower limit of its operatingrange. This low flow operation is theroot cause of many pump reliabilityproblems.

Good communication during thesizing process can help avoid thisand other sizing errors. The plant’srotating machinery group should bebrought in early to work with theprocess designers. Working together,they can explore options other thanpump oversizing. The rotating equip-ment group also can bring pumpmanufacturer expertise to the sizingexercise. Pump manufacturers fre-quently offer other choices such asan upgradable pump, variable speeddrive or other means to improveoperational flexibility.

SITE/INSTALLATION DATAIt is important to communicate

the location and type of installationboth internally and to the pump sup-

Communicating Your Pump NeedsPurchasing and installing a new pump requires a team effort between customer and supplier.By John Bertucci

CENTRIFUGAL PUMPS

HANDBOOK

136 The Pump Handbook Series

Page 136: Centrifugal Pumps Handbook

plier. Teamwork in this area is espe-cially vital in vertical pit pumpinstallations such as cooling towerpumps. Errors in designing the pitand suction approach to the pumpsare easy to correct when the pit con-sists of lines on paper. But they areextremely expensive to correct onceconcrete is poured.

USER RESTRICTIONS ANDPREFERENCES

Every plant has certain thingssuch as seal type, maximum suctionspecific speed and bearing type thatthey like and dislike in their pumps.These preferences and restrictionsare usually based on years of experi-

ence in solving that plant’s particularpump problems. They must beexplained well internally so that thepump data sheet accurately conveysthem to the suppliers. Unfortunate-ly, these preferences usually are inthe heads of the plant rotating equip-ment group while a lot of the pumpselection is done by a project engi-neering group or outside engineeringdesign firm. One way to assure thatthese preferences are given due con-sideration is to require review of thepump data sheet by the rotatingequipment group. Another good wayis to place a member of the rotatingequipment group on the projectdesign team where his/her knowl-edge can be tapped by the projectdesign engineers. A third way ofaccomplishing this is to develop alocal specification that contains thevarious preferences and restrictionsfrom the rotating equipment group.

COMMUNICATION WITH THE PUMP SUPPLIER

Communication with the pumpsupplier should be a two-way street.The information should flow freelyback and forth between the user andsupplier. This should happen even insituations where competitive biddingwill determine the ultimate supplierof a new pump. The only differencein the competitive bid situation isthat all of the data sent from the userto the supplier should go to all sup-pliers equally, with no favoritismshown. Of course, all commercialinformation (prices, delivery details,etc.) should be kept confidential.

INFORMATION OBTAINED FROM PUMP SUPPLIER

Much information can beobtained from the pump supplierthat will aid in future pump mainte-nance and troubleshooting. Some ofthe less obvious or often forgottenitems are listed here.

TROUBLESHOOTING ANDMAINTENANCE

Table 1 shows information thatcan be obtained from the pump’smanufacturer for future use in trou-bleshooting. This is information thatthe manufacturer generally has andwill provide upon request. Table 2lists pump supplier information thatcan be very useful in future pumpmaintenance.

INFORMATION ITEMS POTENTIAL USESSuction Specific Speed 1. Determining Stable Flow Range.

2. Re-rates, especially to lower rates.

Number of Impeller Vanes Vibration Analysis, especially vibrationscaused by low flow.

Seal Flush Flow Rate Calculations and 1. Determine the cause of seal failures.Stuffing Box Pressure 2. Change seal design.

Stable Flow Range 1. Determine the minimum or maximumallowable flow.

2. Determine if pump is in a low or high flow condition.

Thermal Growth 1. Determine if pump is distorting due to thermal forces.

2. Aid in getting good alignment.

Table 1

INFORMATION ITEMS POTENTIAL USESWear Ring Clearances 1. Checking existing clearances to determine if wear

ring replacement is necessary.2. Determine if performance problem was due to

excessive wear ring clearance.

Bearing Number 1. Vibration Analysis - determine ball pass frequencies.(Rolling Element Bearings) 2. Analyze potential bearing upgrades.

Bearing Clearance 1. Analyze vibration problems.(Hydrodynamic Bearings) 2. Set alarm limits on probe type vibration monitors.

3. Determine bearing replacement needs.

Stuffing Box Dimensions Mechanical seal upgrades and changes.

Materials of Construction 1. Determine repair methods.2. Emergency fabrication of replacement parts.

Table 2

DOCUMENT REVIEW BEFORE COPY AFTERPURCHASE PURCHASE

Installation, Operation & Maintenance Manual NO YESCross Sectional Drawing with Parts Identified YES YESDimensional Outline Drawing YES YESSpare Parts List YES YESPerformance Curve YES YESCurve Family YES NOCompleted Data Sheet YES YESTest Data (if applicable) NO YESDriver Data YES YES

Table 3

The Pump Handbook Series 137

Page 137: Centrifugal Pumps Handbook

OTHER DOCUMENTATIONTable 3 shows what can be con-

sidered a good minimum require-ment for a documentation package.

PERFORMANCE IMPROVEMENTSPump performance improve-

ments generally come in two flavors:upgrade of an existing pump that isperforming well, and correction of aproblem pump. In either case, thepump’s original supplier can be anextremely valuable resource.

Be Open to New Ideas. Whenyour friendly pump supplier callsand asks for an appointment, maketime for him (even if it’s not lunchtime). Suppliers are constantly com-ing up with new and better ways todo things. Examples include A and Bgap modifications, new overlaymaterials for severe service and newimpeller designs. You can’t considerthese and other potential solutions toyour problems if you don’t take timeto learn about them from theexperts.

Invite Your Supplier to Par-ticipate. The manufacturer’s localrepresentative should be a regularpart of your maintenance resources.You may have a problem that yourcompetition solved years ago. You’llnever find this out from the competi-tion, but the pump supplier mayalready know the solution. However,

he can’t help if he doesn’t knowabout your problem. In addition, thepump supplier will be intimatelyfamiliar with the design issues of aspecific pump.

Visit Supplier’s Repair Facili-ty or Factory. A visit to the suppli-er’s repair shop can give you muchvaluable information: It enables youto evaluate the shop’s capabilities, incase you ever need them.

It allows you to develop face-to-face contacts with the people whorepair these pumps every day. Agood relationship with these folksmay help you in the future when youneed information. Also, the informa-tion you get there is not filteredthrough a salesman.

A visit to the factory can be evenbetter. It allows you to meet with thepeople who know the most about thedesign of your pump. Factory contactscan get you information in a hurry.They can also help expedite shipmentof desperately needed parts.

Go to Conferences. There aremany good conferences and symposiathat have pumps as a primary theme.One highly recommended conferenceis the International Pump Users Sym-posium held in early March each yearin Houston. Sponsored by TexasA&M University’s TurbomachineryLaboratory, it includes short courses,tutorials, discussion groups and hun-

dreds of pump manufacturer andrelated industry displays. These con-ferences are like having many pumpfactories under one roof. You can gofrom booth to booth and meet repre-sentatives of many companies. Thediscussion groups are also a good wayto interface with other users as wellas suppliers.

Share Knowledge/Experiencewith the Pump Supplier. Usersaccrue valuable practical experiencewith pumps over their years of opera-tion. This experience should beshared with the pump supplier. It willhelp the supplier improve his prod-ucts and eventually benefit all users.Also, it is only fair that knowledge goboth ways in any relationship.

In summary, your pump suppli-er will never know or understandyour needs if you don’t take the timeand effort to develop a mutually ben-eficial relationship. Most pump sup-pliers place a high priority onmeeting their customers’ needs, butthey need all the help that they canget. ■

John Bertucci is a MechanicalEquipment Engineer with the NorcoManufacturing complex of Shell OilProducts Co.

138 The Pump Handbook Series

Page 138: Centrifugal Pumps Handbook

All centrifugal pumpshave two major com-ponents: the rotatingelement and the pumpcase. Together theyestablish how muchhead will be generat-ed, best efficiency

point (BEP) capacity, the slope of thehead capacity curve and net positivesuction head required (NPSHR), Fig-ure 1. The rotating element consistsof a shaft and one or more impellerswhose function is to convert mechan-ical energy into high velocity kineticenergy. The pump case directs liquidto the impeller from the suction noz-zle, collects the liquid dischargingfrom the impeller, and then convertsthe kinetic energy into pressure bycontrolled deceleration in the diffu-sion chamber immediately following

the volute throat.

TYPES OF VOLUTESSingle volute pumps have only

one volute throat and one diffusionchamber, and because of their sim-ple casting geometry, they can beproduced at lower cost than themore complex double volute designs.Figure 2 shows a single stage versionwith the diffusion leading directly topump discharge. Figure 3 is a multi-stage version in which the liquid,following diffusion, is directedthrough the crossover to the nextstage impeller. Pressure distributionaround the impeller is non-uniformand produces a radial load on theimpeller which, depending on thedeveloped head, may deflect thepump shaft and cause wear atimpeller wear rings and seal faces.

Single volutes are used routinely

in slurry and sewage pumps to mini-mize plugging at the throat, and onlow head pumps where radial loadsare nominal. They are also used onlow specific speed pumps where thethroat area and hydraulic passagesare too small to cast as a doublevolute.

Double volute casings wereintroduced to minimize the radialthrust problems of single volutepumps. They are actually two singlevolute designs 180° apart with a totalthroat area equivalent to a compara-ble single volute design. The non-uniform pressure distributions areopposed, thereby greatly reducingradial loads (Figure 4). Double volutepumps are the preferred choice onhigher head pumps.

TYPES OF IMPELLERSPumps can be built with a single

impeller or, in the case of high pres-sure applications, with two or moreimpellers. They will be either singleentry or double entry type, morecommonly known as single or dou-ble suction (Figure 5). Because dou-ble suction impellers have a greatertotal eye area, velocity of the liquidentering the eye is reduced, produc-ing a lower NPSHR. The shape of theimpeller depends on specific speed(Ns), which should only be calculat-ed at BEP with maximum diameterimpeller. In U.S. units this is:

Ns=RPM x GPM.5

(Head/Stage feet).75

Figure 6 shows the change inshape from the low Ns radial flowimpellers to the high Ns axial flowtypes.

Whereas the suction geometry isselected to reduce inlet losses for low

Impellers and Volutes:Power with Control

A review of centrifugal pump impeller and volute design applications can help you optimize both power and control.

By Robert R. Ross

0 400 800 1200 1600 2000 2400 2800

550

500

450

400

350

300

250

200

150

100

6050

40

30

400

300

200

100

0

90

80

70

60

50

40

30

20

10

0

HEAD

IN F

EET

NPSH

R FT

BRAK

E H.

P.

EFFI

CIEN

CY %

B.E.

P.

3HP SP. GR. 1.0

HEAD-CAP

E.F.F.

%

NPSHR

CAPACITY GPMPump performance curve

THE PUMP HANDBOOK SERIES 139

CENTRIFUGAL PUMPS

HANDBOOK

FIGURE 1

Page 139: Centrifugal Pumps Handbook

NPSHR, the discharge geometry isselected to satisfy the required head,slope of the head capacity curve andBEP capacity.

HEAD CAPACITY CURVE SLOPEANALYSIS

The desired slope is determinedduring system analysis with the per-centage rise to shutoff (zero flow)from rated head often determined bysystem limits. When the pump isstarted against a closed dischargevalve, the pressure up to the valvewill be the pump differential head at shutoff plus suction pressure.Because this should not exceed thesafe working pressure of the system,the rise to shutoff can be critical andis controlled by impeller dischargegeometry.

An evaluation of the system headcurve is needed to determine if thepump should have a flat or a steepcurve. This is a graphical plot of thetotal static head and friction lossesfor various flow rates. For anydesired flow rate, the head to be gen-erated by the pump is at the intersec-tion of the head capacity curve andsystem head curve. In a simple pumpapplication in which system head isdue entirely to friction loss, a flathead capacity curve with 10 to 20%rise to shutoff from the head at ratedcapacity would satisfy the applicationand minimize shutoff pressure on thesystem (Figure 7). Where systemhead consists of both friction and sta-tic head — that is, where there is achange in elevation — a flat curvealso would be appropriate if little or

no change was anticipated in the sta-tic head (Figure 8).

Figure 9 represents an applica-tion in which changes in static headcaused by changes in the suctiontank level result in a range of systemhead curves. For illustration, twopump head capacity curves havebeen superimposed — one flat, theother steep. The advantages and dis-advantages of both must be consid-ered in deciding which pump curveis more suitable for the application.

Advantages of the flat curve arelow shutoff pressure and relativelysmall differences in operating pres-sure as the system head moves. Thedisadvantage is a larger variation inflow rates.

Advantages of the steep curve aresmaller variations in flow rates andadditional head margin to accommo-date potential increases in static head.Disadvantages are high shutoff pres-sure and larger variations in head.

IMPELLER DISCHARGE GEOMETRYVarious methods are used to

modify the impeller so the percent-age rise to shutoff will match theslope resulting from system headanalysis. Among these are changes inthe number of vanes, changes in thevane discharge angle and changes inthe exit width b2. The curve can bechanged with variations in vanenumber and discharge angle, whileBEP capacity is held constant bychanging b2. Another method is touse a constant discharge angle and b2with changes in the number of vanesonly.

SINGLE SUCTIONIMPELLER

DOUBLE SUCTIONIMPELLER

Centrifugal impellers

Volute Throat

Impeller

Diffusion Chamber

RadialFlow

Semi-RadialFlow

MixedFlow

AxialFlow

1000 3000 6000 12000

SPECIFIC SPEED Ns

FIGURE 2

FIGURE 3

FIGURE 4

FIGURE 5

FIGURE 6

140 The Pump Handbook Series

Page 140: Centrifugal Pumps Handbook

BEP CAPACITYThe impeller discharge geometry

and volute throat area establish BEPcapacity. By adjusting the ratio of liq-uid velocity leaving the impeller to liq-uid velocity entering the volute throat,

BEP can be increased or decreased.Modifications of this type are used toupsize or downsize existing pumpshydraulically, moving BEP to the nor-mal operating capacity for optimumefficiency and generating significantsavings in the cost of power. ■

Robert R. Ross is Director of Engi-neering for BW/IP International and amember of the Pumps and SystemsEditorial Advisory Team.

Editor’s Note: Some text and fig-ures for this article have been excerptedwith publisher’s permission fromLobanoff, V. S. and Ross, R.R., Cen-trifugal Pumps: Design and Applica-tion, 2nd Edition, Gulf PublishingCompany, Houston, TX 1992.

SUCTION

TANKLOW

NORMALOPERATIO

N

SUCTION TANKFULL

FLAT

STEEP

EVALUATING PUMP CURVES AGAINSTVARIABLE SYSTEM HEAD CURVES

Capacity

Head

SYSTEM HEAD CURVE, ALL FRICTION

Capacity

SYSTEM HEAD CURVE, STATIC HEAD PLUS FRICTION

Capacity

FIGURE 7

FIGURE 8

FIGURE 9

THE PUMP HANDBOOK SERIES 141

Page 141: Centrifugal Pumps Handbook

The Fluid CatalyticCracking Unit (FCCU)is one of severalprocesses critical to arefinery’s productivity.Its long term, safe oper-ation translates intoincreased production

and profitability. Yet prior to intro-ducing fully lined pump technologyfor use in FCC (Fluid CatalyticCracking) main column bottomsapplications, refineries replaced orrepaired conventional double voluteAPI process pumps several times ayear. This caused serious safety risksand a process shutdown frequencyunacceptable in today’s productionenvironment.

The fully lined slurry pump hasthus emerged as the pump technolo-gy of choice for providing 3 – 5 yearsof maintenance-free operation forFCCU refinery bottoms applications.This particular application involvespumping a highly erosive high tem-perature (350 – 800°F) slurry atflows to 12,000 gpm, pressures to600 psig and heads from 90 – 900feet. Catalysts used in FCC processesare also extremely erosive, and theyare applied in varying concentrationsdepending on the process unit con-figuration and/or upset operatingconditions.

The fully lined slurry pumpdesign (Figure 1) is an engineeredapproach to providing long term, reli-able pump performance in this severe-duty application. To understand howthis new pump technology meets suchdemanding requirements, we will takea closer look at FCC processes and thedesign of fully lined slurry pumps.

FCCU PROCESSES

Conventional Fluid CatalyticCracking. A typical fluid catalyticcracking unit (Figure 2) consists of areactor, catalyst regenerator andfractionator column. This processconverts straight run heavy gas oilfrom the crude distilling unit, andflasher tops from the vacuum flasherunit, into high octane gasoline, lightfuel oils and olefin-rich light gases.

In the vertical reactor vessel,vaporized oil contacts fluidized cata-lyst particles, causing a reaction thatyields lighter hydrocarbon and coke.During the reaction, carbon (coke) isdeposited on the catalyst, rendering it

inactive. This inactive catalyst is recir-culated from the cyclones at the top ofthe reactor back to the regeneratorwhere the coke is combusted, rejuve-nating the catalyst. Vaporized crackedproducts flow through the cyclones atthe top of the reactor into the vaporline that feeds the bottom of the mainfractionator column. The cyclonesoperate at less than 100% efficiency sothat some coke and catalyst particlescontinuously reach the fractionator.

In this type of FCC process, themain column bottoms pumps mustpump the bottom oil at a high ratethrough the heat exchanger and over

Fully Lined Slurry Pump for FCCU Bottoms Use

By replacing conventional API pumps with fully lined slurry pumps in FCCUbottoms applications, refineries are improving production and profitability.

By Dan Clark and Julio R. Cayro

FIGURE 1

142 The Pump Handbook Series

CENTRIFUGAL PUMPS

HANDBOOK

Fully lined slurry pump design

Page 142: Centrifugal Pumps Handbook

a vapor contact section within thefractionator tower. This desuperheatsand scrubs the fine particles of cata-lyst from the reactor vapors enteringthe fractionator without causing oilcoking. A small concentration of alu-mina-based catalyst particles thathave been scrubbed out of the vaporsis continuously circulated with themain bottoms product, causing ero-sion of the pump internals.

New FCC Processes. Unlike con-ventional FCC processes that recyclepart of the main column bottomsdirectly back to the reactor vessel,newer FCC processes such as theUOP process continually circulatethe main column bottoms in a closedloop through heat exchangers andback to the fractionator tower. Withthe closed loop design, catalyst con-centrations run as high as 2% byweight, compared to conventionalsystems where the catalyst concen-tration is 0.25 – 0.5%.

With conventional FCC process-es, catalyst levels reached 1 – 1.5%only during prolonged upset operat-ing conditions, destroying conven-tional API process pump internals ina matter of days or weeks (Photo 1).Today, a 2% concentration of cata-lyst continuously circulating withthe main column bottoms is consid-ered a normal operating condition.

FULLY LINED SLURRY PUMPTECHNOLOGY

Unlike conventional API processpumps designed for maximum effi-ciency with clean liquids, fully linedslurry pumps are engineered to pro-vide maximum reliability when han-dling abrasive hydrocarbon slurries.All pumps are selected to operate inthe optimum hydraulic fit (80 –110% BEP) for specified flow, headand erosive characteristics of theslurry. Design considerations such asusing larger diameter impellers atlower speeds (870 – 1,770 rpm),selecting proper construction materi-als and maximizing appropriatemechanical seal designs all help opti-mize life cycle cost.

Liners. The fully lined slurrypump uses replaceable, abrasion-resistant 28% chrome iron liners toprotect the pressure casing, providing5 – 6 times the life of diffusion coatedCA6nm components (Figure 3).

Abrasion-resistant liners aremachined, toleranced componentsthat form the hydraulic wet end of the

pump. They are easily replaced indi-vidually or as a set. Flow stream tur-bulence is reduced by using a 125 rmsmachine finish on the liner surfaces.

Rotating Elements. Like the lin-ers, impellers are constructed ofabrasion resistant 28% chrome iron.These high efficiency enclosedimpellers (Photo 2) use front andback repelling vanes to reduce slurryrecirculation. Repelling vanes elimi-nate the clean oil flush required by

conventional wear ring impellerdesigns. Large open passages reducefrictional losses and allow maximumsolids-passing capability to avoidclogging due to coke buildup in thefractionator column.

Several impeller mounting con-figurations are available dependingon horsepower and catalyst slurryproperties. Impellers can be fastenedto the shaft using a tapered polygonor straight bore. Each is locked in

Conventional API process pump internals can be destroyed in a matterof days during prolonged upset operating conditions.

PHOT

O CO

URTE

SY O

F LA

WRE

NCE

PUM

PS

The Pump Handbook Series 143

PHOTO 1

FIGURE 2

PRESSURE REDUCING CHAMBER

FLUE GAS TO

CO BOILER

GAS & GASOLINETO

GAS-CONCENTRATIONPLANT

HEAVY-CYCLE GAS OIL

CLARIFIED OIL

SLURRY SETTLER

REGENERATOR

LIGHT-CYCLE GAS OIL

FRACTIONATOR

REACTOR

AIR

CHARGE

CATALYSTSTRIPPER

STEAM

UOP “Straight-Riser” Fluid Catalytic Cracking Unit

Page 143: Centrifugal Pumps Handbook

place with an enclosed impeller nutand then secured with a locking bolt.

As shown in Figure 1, fullylined slurry pumps use large diame-ter shafts that meet the stiffness cri-teria set by API-610. Each shaft isengineered to provide space for sin-gle, double or tandem seal arrange-ments and a conservative L3/D4deflection index. Stiff shaft designslimit deflection, maximize mechani-cal seal and bearing life, and mini-mize vibration.

A heavy duty bearing assemblyemploys 7300 Series bearings withslight preloads to support the shaft.The anti-friction bearings provide aminimum L-10 rated life of 100,000hours at the rated pump condition.Thrust bearings are duplex, angularcontact type mounted back to back.The radial bearing is either the anti-friction ball or spherical rollerarrangement, depending on radialloads and rotative speeds.

Bearings are lubricated usingring oil, flood oil, oil mist and forcedfeed arrangements. Oil mist is usedto minimize friction, but it is notacceptable for cooling bearings thatare subject to heat transferred froman external source such as the shaft.Thrust and radial bearing covers areequipped with isolators that have adeep grooved labyrinth which pre-vents oil from escaping from thebearing frame.

MECHANICAL SEAL CONFIGURATIONSAND OPTIONS

All fully lined slurry pumpsoperating in FCCU bottoms applica-tions feature a removable seal cham-ber that allows pressure testing of theseal before installation. This designgives the user the ability to switchseal types easily to accommodate cat-alyst slurry changes and/or to complywith environmental regulations.

Seal chambers are designed toaccommodate single, tandem anddouble mechanical seal arrange-ments. A replaceable throat bushingmechanically fastened to the sealchamber rides on the impeller hub.The shaft sleeve is 316 stainlesssteel, and its straight design allowseasy adjustment of impeller clear-ances without disturbing the seal set-ting. A Ringfeder locking collarreplaces the conventional lockingcollar with set screws. The Ringfed-er eliminates

the need for set screws to drive theshaft sleeve. It also seals the cavitybetween the shaft and sleeve bycompressing the sleeve onto theshaft. This clamping arrangementprevents slippage of the sleeve onthe shaft under severe slurry upsetconditions, and it eliminates theneed for grafoil packing by providing

a high temperature, high pressuremetal-to-metal seal between theshaft and sleeve.

STARTUP AND OPERATION ISSUESThe startup procedure for fully

lined slurry pumps operating in theFCCU bottoms application is criticalto the pump’s long term performance.

Enclosed impeller with front and back repelling vanes to reduce circula-tion of the slurry.

PHOT

O CO

URTE

SY O

F LA

WRE

NCE

PUM

PS

FIGURE 3

PHOTO 2

144 The Pump Handbook Series

COATED API PUMP

0 6 12 18 24 30 36 42 48 54 60 66 72

500

450

400

350

300

250

200

150

100

50

0

FULLY LINED API PUMP

Page 144: Centrifugal Pumps Handbook

To avoid thermally shocking thehard metal liners and impeller, hotoil is injected into the pump casing topreheat these internal componentsgradually — at a rate not exceeding150°F per hour.

Initially, the oil steam is intro-duced at less than 250°F througheither the casing drain or seal flushconnection at a pressure higher thanthe downstream discharge pressure..The steam is then allowed to flowthrough the casing. Operators areasked to maintain a constant preheatrate until the pump is heated to with-in 150°F of its actual operating tem-perature and allowed to soak for onehour at the maximum preheat tem-perature before startup.

Providing the proper cooling tothe bearing frames and pedestals,and flush oil to the mechanicalseals, becomes very important oncethe pump reaches its startup tem-perature.

To ensure long term successfuloperation of a fully lined slurry pumpin the FCCU bottoms application, thefollowing steps are recommended.• Provide dual strainers in front ofthe pump to allow cleaning thestrainer without shutting down thepump.• Use slurry impeller designs toallow for larger mesh openings in thesuction strainers and increase thecycle time between cleaning by 300%.• Equip the FCCU system with aminimum flow bypass line so the

pump can be operated continuouslyat its BEP. • Use mechanical seals in the car-tridge canister arrangement to allowperformance testing during sealdesign stages and hydrotestingbefore installing the seal. • Consider using double seal config-urations for added safety and pro-vide for inboard seal faces to runcontinually on a clean liquid.

MAINTENANCE ISSUESFully lined slurry pumps operat-

ing in the FCCU bottoms applicationrequire maintenance, repair orreplacement of the casing liners andimpeller every 3 – 5 years. Thismaintenance is simplified by a backpull-out arrangement allowing accessto the liners and impeller without dis-turbing suction and discharge piping.

After assembly, impeller clear-ances can be adjusted easily by mov-ing the thrust bearing cartridgerelative to the bearing frame. Shimsare then placed between the car-tridge and bearing frame to lock theshaft and impeller into position. Thelarge oil reservoir for the bearingsensures a continuous source of cleanoil to lubricate the bearings. Oil levelis monitored using a 2” bulls eye inthe side of the bearing frame.

The entire pump can be rebuiltby mechanical craftsmen using stan-dard shop tools. All wear compo-nents are replaceable, and all fits andclearances are standard.

LONG TERM PERFORMANCESeveral hundred fully lined slur-

ry pumps are operating in FCCU bot-toms applications in refineriesaround the world. More specifically,a fully lined slurry pump installed ina UOP FCCU unit has been runningcontinuously since 1991, requiringonly scheduled maintenance.

Fully lined slurry pump tech-nologies continue to evolve to meetthe increasingly demanding perfor-mance criteria of the FCCU bottomsapplication. Refineries that haveinstalled fully lined slurry pumps intheir cracking units have eliminatedthe expense of replacing convention-al API process pumps several times ayear. More important, they’ve elimi-nated the safety risk of catastrophicpump failure and the prohibitive costof shutting down the FCCU for sev-eral days. ■

Dan Clark has been withLawrence Pumps Inc. since 1973 andhas overseen the development of theFCCU bottoms fully lined slurry pump.

Julio R. Cayro is a MechanicalEquipment Consultant and owner ofCayro Engineering Company in Hous-ton, TX.

The Pump Handbook Series 145

Page 145: Centrifugal Pumps Handbook

In recent years, the case for usingsealless centrifugal pumps hascentered mainly on zero emis-sions—and the fact that they do

not require seal support systems andperiodic mechanical seal replace-ment.

But this is only part of the story.Design modifications and accessoriesare expanding the performanceranges of both canned motor andmagnetic drive pumps. Here is ageneric comparison of what thesepumps offer when it comes tohydraulic application features andoptions:

1. DOUBLE CONTAINMENTThis is an important considera-

tion in services that pose extremehealth and/or safety concerns. Several manufacturers of cannedmotor pumps (CMP’s) offer double-containment (i.e. welded primarycontainment, hermetically sealedsecondary containment). Because ofefficiency issues, few magnetic dri-ve pump (MDP’s) manufacturershave approached secondary con-tainment.

Most MDP’s can provide sec-ondary “control” by utilizingmechanical seals on the OMR shaftpenetration. Additionally, an MDPprovides a thicker containmentshell, and thus more resistance topenetration by corrosive or mechan-ical failure. Typical CMP primaryliner thicknesses are 0.022–0.35″,while an MDP containment shell is0.029–0.060″. Bearing monitoringplays a role, however, because bear-ing and internal rotor positions areeasier to monitor in CMP’s, bydesign, than in MDP’s. This enablesthe CMP to detect extreme bearingwear prior to containment shell con-tact or violation.

2. SOLIDS/SLURRY HANDLINGBoth the canned motor and

magnetic drive designs will handlemoderate amounts of solids, andoptional designs for both will handlehigher concentrations of slurries.For CMP’s, if low concentrations of

solids are present (about 2%) andtheir size is relatively small (25µmmaximum), hard bearings runningagainst a hardened journal workwell. CMP’s can handle increasedsolids if they are outfitted withexternal flush or filters to removeparticulates from the pumpagebefore they circulate around thebearings. In fact, CMP’s with a slur-ry modification are able to handleslurries in the concentration rangeshandled by a standard centrifugalpump. An MDP cannot be isolatedas easily as the CMP, and thereforewould require a completely differ-ent internal bearing-mag couplingflow path than is conventionallyoffered.

3. HEAT INPUTBoth sealless centrifugal pump

designs add heat by hydraulic anddrive inefficiencies. For a CMPdesign a “high efficiency” motor willbe 80–85% efficient. But because ofthe ease in isolating the motor area,CMP’s can offer optional configura-tions to control the fluid tempera-ture, pressure, or both, to preventproduct vaporization. However,“heat soak” can occur. In this situa-tion the process fluid will be heatedto higher temperatures at potentiallylower pressures than during normaloperation. This is because a CMPmotor tends to be a large insulatedmass once the unit is shut down.This may result in flashing of thecontained fluid with a potential ofvapor locking if restarted.

4. COOLING REQUIREMENTSBoth the CMP and MDP can be

configured for operation at high tem-peratures. However, permanentmagnets can tolerate heat better thanmotor windings can, so MDP’s areable to pump hot liquids—up to750°F—with just air cooling. Cannedmotor pumps require water coolingjackets for high temperature service.For example, some CMP designs canoperate up to 1,000°F with watercooling. Various designs require tem-perature limitations by component

and must be evaluated on a case-by-case basis.

Several things must be consid-ered on MDP designs. A synchro-nous drive is more efficient than aneddy current drive. A non-metalliccontainment shell is more efficientthan a metallic shell. The MDP lacksthe large insulated mass around thecontainment shell and is less suscep-tible to “heat soak.”

5. JACKETINGEither design can accept steam

or hot oil jackets added to the pumpto maintain the proper temperatureof the product with a high freezingpoint. This insures that the pumpageremains a liquid during pump opera-tion and shut-down.

CMP designs allow jacketing ofthe pump case, stator and rear bear-ing housing. Users must be carefulnot to exceed the thermal limits ofthe stator insulation with the heatingmedia. Most MDP’s can jacket thecasing and add some heat in the areaof the containment shell without ful-ly encapsulating it.

6. HIGH SUCTION PRESSURETypically, the CMP is more effi-

cient than the MDP in high pressureapplications because of the increasedprimary containment thickness. Acanned motor pump is a pressurevessel since the stator windings lendadditional mechanical strength.

For applications in which thesuction pressure and maximumallowable working pressure require-

Options for Sealless Centrifugals

Cutaway of a Kontro A-range ANSI seal-less magnet drive pump used in chemicalprocessing services.

PHOT

O CO

URTE

SY S

UNST

RAND

FLU

ID H

ANDL

ING

CENTRIFUGAL PUMPS

HANDBOOK

146 The Pump Handbook Series

Page 146: Centrifugal Pumps Handbook

ments exceed the standard pressuredesign capability of either a CMP orMDP, optional designs are availablefor both. In fact, some CMP’s areavailable that can have as high as5,000 psi system pressure design.

Modifications in the CMP designfor high pressure applicationsinclude the use of primary contain-ment shell backing rings, thicker sec-ondary containment shells,additional pressure-containing bolt-ing and high pressure terminalplates. For MDP’s high pressureapplication modifications includeusage of a thicker containment shelland additional bolting.

In addition to application fea-tures, users should be aware of certainhydraulic and design differencesbetween canned motor and magneticdrive pumps. Here again, both offeradvantages.

7. ANSIWhile a few manufacturers of

canned motor pumps offer units withANSI dimensions and/or hydraulics,most do not. On the other hand, themajority of mag drive pump suppliersdo offer ANSI dimensions andhydraulics. In addition, most manu-facturers of sealed and sealless ANSIpumps offer interchangeabilitybetween their pumps’ wet ends andbearing frames.

8. EFFICIENCYCMP wire-to-water efficiency is

defined as hydraulic times motor effi-ciency. As mentioned in the sectionon heat input, a typical CMP motorwill be 80–85% efficient. MDP wire-to-water efficiency is defined ashydraulic efficiency times motor effi-ciency times coupling efficiency.Again, the magnetic coupling is80–85% efficient. However, contain-ment shell metallurgy (or lack there-of) and magnetic coupling type play abig role in coupling efficiency.

Further complicating the effi-ciency discussion is wet end hard-ware. Impeller and casing geometryplay a vital role in hydraulic effi-ciency.

For example, a Barske design(open radial blade impeller and dif-fuser discharge) is usually more effi-

cient in low flow/high headhydraulics. A Francis design impeller(enclosed with backswept vanes andan increasing-radius volute) is moreefficient at moderate to high flowswith low to medium heads.

You need to evaluate “wire-to-water” efficiency to get a truly accu-rate picture of efficiency.

9. INTERNAL CLEARANCESWhile clearances between bearing

ID and mating surfaces are typicallythe same (0.003–0.007″), differencesoccur between other rotating parts.

Typical CMP clearances betweenthe rotor and stator liners vary bymanufacturer between 0.018–0.044″radially. Typical MDP clearancesbetween the inner magnetic ring(MR) and containment shell rangefrom 0.030–0.045″ radially. This larg-er clearance gives MDP’s the advan-tage of allowing more bearing wear tooccur prior to containment shell con-tact.

10. BEARING MONITORINGWhile different monitoring meth-

ods are available for both MDP’s andCMP’s, the latter design lends itselfmore to real bearing monitoring. ACMP bearing monitor can provideaxial, radial and liner corrosive wearindications. However, bearing moni-

toring features vary by manufacturerand care must be exercised whenselecting a sealless pump vendor toinsure that the desired monitoringfeatures are provided.

11. SPACE CONSIDERATIONSIn general, a CMP (integral pump

and motor) occupies less of a footprint than a comparable MDP (pump,coupling and motor). However, close-coupled MDP’s are available, andthese may require the same or lessspace than a CMP.

12. COUPLING ALIGNMENT ANDVENTING

Coupling alignment is not requiredfor CMP installations because thepump impeller is directly mounted onthe motor shaft inside of the contain-ment area, so no coupling exists.

Typical MDP installations utilizeframe-mounted motors which requirecoupling alignment. Some MDP sup-pliers, however, offer close-coupleddesigns which eliminate couplingalignment. Both MDP and CMPdesigns are usually self-venting backinto the process piping and do notrequire additional external lines.

In addition to the issues alreadydiscussed, some manufacturers offercanned motor and magnetic drive cen-trifugal pumps with options such as thefollowing: two-phase flow designs;tachometers; open impellers (non-shroud) with isolated motor sections forsolids handling; diagnostics that indi-cate rotor position, stator liner rupture,temperature and pressure; vibrationpads, and redundant systems to indi-cate breach and contain fluid.

The best thing you can do ifyou’re considering options for yoursealless centrifugal pumps is to makesure your supplier(s) know everythingabout your application, particularlytemperatures and vapor pressure atstartup and shutdown, not just normaloperating conditions. ■

This article was developed with theassistance of Steven A. Jaskiewicz, ofCrane Chempump (Warrington, PA)and David Carr, of Sundstrand FluidHandling (Arvada, CO).

Cutaway view of a canned motor pump.These pumps are used on a wide variety offluids at temperatures from cryogenic ser-vice to 1000° F and at system pressuresup to 5,000 PSI.

PHOT

O CO

URTE

SY C

RANE

CO.

, CHE

MPU

MP

DIVI

SION

The Pump Handbook Series 147

Page 147: Centrifugal Pumps Handbook

Arevolution swept through thechemical process pumpindustry more than 30 yearsago. It was the beginning ofchemical pump design anddimension standardization.

Before the 1960s, chemical processpump manufacturers offered a prolifer-ation of designs. Each manufacturer hadits own design and dimensional enve-lope. Industrial users faced significantpiping, baseplate design and potentialfoundation changes if existing pumpshad to be replaced. This very expensivepossibility became the driving forcebehind the development of what indus-try today refers to as an "ANSI" pump.During the 1960s and 1970s, the Ameri-can Voluntary Standard (AVS pump)served as the chemical process pumpstandard. In 1974, the American Nation-al Standards Institute (ANSI) used theAVS standard as the foundation for itsB73.1 specification covering chemicalprocess pumps.

After several revisions today’sANSI/ASME B73.1M - 1991 Specifica-tion for Horizontal End Suction Cen-trifugal Pumps for Chemical Processserves as the industry standard. Itcovers dimensional interchangeabili-ty requirements for 20 pump sizes.This includes mounting dimensions,suction and discharge flange size andlocation, input shaft size, baseplatesand foundation bolt holes. It alsoaddresses many mechanical designfeatures such as pressure limits, tem-perature limits, drain and gauge connections and seal chamber dimen-sions. This enables today’s pump userto replace a pump with one from adifferent manufacturer easily.

Chemical process pump specifi-cations developed through the B73committee include:

• ANSI/ASME B73.1M - 1991,Specification for Horizontal End Suc-tion Chemical Pumps for ChemicalProcess (Photo 1)

• ANSI/ASME B73.2M - 1991,Specification for Vertical In-Line Cen-trifugal Pumps for Chemical Process(Photo 2)

• ANSI/ASME B73.3M - (inprocess), Specification for SeallessHorizontal End Suction CentrifugalPumps for Chemical Process

• ANSI/ASME B73.5M - (inprocess), Specification for Thermo-plastic and Thermoset Polymer Mate-rial Horizontal End SuctionCentrifugal Pumps for ChemicalProcess

PRINCIPLES OF OPERATIONThe pumps covered by ANSI/

ASME B73.1M are classified as endsuction centrifugal pumps. Centrifu-gal pumps use an impeller and angledvanes to impart velocity to the liquidentering the pump. The liquid leavingthe impeller is collected by the pumpcasing, an action that converts a por-tion of the fluid velocity into pres-sure. Either mechanical packing in astuffing box or a mechanical seal in aseal chamber is used to seal the rotat-ing shaft. Figure 1 shows the basiccomponents of an ANSI pump.

Impellers. Centrifugal impellerdesigns are of two basic types, theopen style and the closed style, asshown in Figure 2. Most ANSI pumpsemploy some type of open impellerand axial adjustment feature. Thisallows critical operating clearanceswithin the pump to be maintained,which is important for maximizinghydraulic performance. The openimpeller is also better at handlingsolids and pumping liquids with

entrained gases. Open impellers –particularly those running at 3600rpm – must be carefully engineeredto control axial thrust, seal chamberpressure and mechanical integrity.Closed impellers are typicallyemployed in less corrosive environ-

Tips for Selecting ANSI Process Pumps

Versatility is the key to ANSI process pump applications.

By Charles Cappellino and Richard Blong

CENTRIFUGAL PUMPS

HANDBOOK

148 The Pump Handbook Series

Photo 1. Example of a horizontal metalANSI process pump

PHOT

O CO

URTE

SY O

F GO

ULDS

PUM

PS, I

NC.

Photo 2. Example of vertical in-line ANSIcentrifugal pump for chemical processservice

PHOT

O CO

URTE

SY O

F GO

ULDS

PUM

PS, I

NC.

Page 148: Centrifugal Pumps Handbook

ments such as light duty chemical,petrochemical and utility applica-tions. This is because closed impellersutilize renewable wear rings to main-tain performance-sensitive runningclearances. Renewable wear rings aresubject to crevice corrosion and aregenerally undesirable for corrosiveservices.

Casings. Three basic types ofcentrifugal pump casing designs areused in chemical process pumps: thecircular volute, single volute and dou-ble volute (Figure 3). The differentcasing designs are used to reducehydraulic radial loads. Pumpsdesigned to operate at extremely lowflows normally use circular volutes tominimize hydraulic radial loads. Sin-gle volute casings are simple to manu-facture and are the most commonlyused for ANSI pumps. Larger pumpsizes require a double volute toreduce hydraulic radial loads. Thecasings utilize ANSI/ASME B16.5Class 150 or Class 300 flanges for suc-tion and discharge connections.

Casings must withstand a hydrostaticpressure test of 1.5 times the maxi-mum design pressure for the materialused, and they must employ an0.125" corrosion allowance by design.

Seal Chambers. ANSI/ASMEB73.1M provides dimensional guide-lines for shaft sealing. The guidelinescover a stuffing box design used forpacking, a large diameter cylindricalseal chamber and a self-venting sealchamber used for mechanical faceseals. Large radial clearance betweenthe shaft and the inside of the sealchamber is specified due to its impor-tance to mechanical seal face tempera-tures. The large radial clearance alsoallows mechanical seal manufacturersto build more robust and reliabledesigns. Because mechanical seals areone of the most significant causes ofpump downtime, most ANSI pumpmanufacturers have developed newseal chamber designs that enhance theoperating environment for mechanicalseals. Typical seal chambers offeredby manufacturers are shown in Figure

4. The most recent self-venting designsincorporate some type of flow modify-ing ribs or superior performance vanesto control solids and entrained gas.Most chemical pump manufacturersoffer several seal chamber designs, aswell as selection guidance for variousservices.

Bearing Housings . ANSI/ASME B73.1M requires a bearingselection that provides 17,500 hoursof life for the radial and axial thrustbearings, calculated according toANSI/AFBMA-9&11. This typicallyresults in a double row thrust bearingbeing used at the coupling end of theshaft and a deep groove ball bearingat the impeller end. (Photo 3)

The pump industry is steadilyimproving the reliability of pumpcomponents. To increase mean timebetween planned maintenance, manymanufacturers have improved bear-ing housing designs and added fea-tures to improve reliability. SomeANSI pump manufacturers offerheavy-duty housings that use angularcontact bearing pairs to handle higherhydraulic thrust loads. Most bearinghousings are sealed using either lip

The Pump Handbook Series 149

Rotation

Figure 1. Basic components of an ANSI B73 pump

Open Impeller Closed Impeller

Figure 2. Centrifugal impeller designs

Circular Volute Single Volute Double Volute

Figure 3. Circular, single and double volute casing designs

Page 149: Centrifugal Pumps Handbook

150 The Pump Handbook Series

seals or higher performance labyrinthseals to prevent lubrication contami-nation–the number one cause of pre-mature bearing failure. Manymanufacturers also have increasedthe capacity of oil sumps to providesuperior heat transfer and cooler run-ning bearings. Most housings havesome type of finned cooler that issubmerged in the oil sump. This con-trols oil temperature in hot services.And large diameter (1") sight glassesare incorporated into many designs toprovide a means of viewing oil condi-tion and level.

Baseplate Designs. ANSI/ASMEB73.1M specifies a set of baseplatedimensions covering motor sizes typi-cally required for the full range ofANSI pump sizes. Dimensions speci-fied include the pump and motormounting surfaces, bedplate footprintand foundation bolt hole locations.Proper baseplate design and installa-tion are necessary to maintain accu-rate pump and driver alignment. Thislengthens the life of bearings andseals, which are sensitive to vibrationand correct alignment.

Proper baseplate selection is the keyto maximizing mean time betweenplanned maintenance. Most pump man-ufacturers offer a selection of baseplatedesigns. Camber top cast iron baseplatesoffer heavy-duty construction withmachined pump and motor pads. Theyalso have good vibration damping char-acteristics. Fabricated steel baseplatesprovide an economical choice in carbonsteel and the option of various metallur-gies such as stainless steel. Many ANSIpump manufacturers also offer sometype of nonmetallic composite baseplate(such as fiberglass reinforced plastic -FRP) for superior corrosion resistance.FRP bases are used with FRP pumps aswell as high alloy pumps. A recent addi-tion to the bedplate options is a heavy-duty fabricated baseplate with integraladjustment features such as adjustmentscrews and baseplate leveling screws(Photo 4). Finally, most bedplates can bestilt or spring mounted. These supportsraise a pump above the floor forimproved cleaning access, and theyaccommodate piping thermal expansion.Stilt and spring-mounted designs mustbe carefully engineered to ensure properrigidity. This will maintain alignmentand avoid vibration problems.

CHEMICAL PROCESS PUMP SELECTION

The ANSI/ASME chemicalprocess pump is the most widelyused centrifugal pump in industry. Itswide use is attributed to its adaptabil-ity to a wide range of process servicepumping conditions. It can be mount-ed vertically to save installationspace, or vertically suspended for usein a sump application. Flexibility incasing and impeller designs enable itto handle extremely low flows, pumpsolids, move highly corrosive liquids,self-prime or withstand temperaturesto 700ºF (371ºC). Recently, the samebasic design has been made sealless byeliminating the need for packing ormechanical seals to seal liquid in thepump. These pumps are magnetic dri-ve ANSI/ ASME chemical processpumps. ANSI/ASME chemical processpumps are perhaps the most versatilepumps in the world.

To help in applying the varioustypes of chemical process pumps, aselection guide is shown in Figure 5.The pump type is shown vertically,and the service parameter is listedhorizontally. ANSI chemical pumpsolutions are available for nearly anypumping service parameter.

Figure 4. Seal chamber stylesA. Standard bore (packed box) is charac-terized by long, narrow cross section.Originally designed for soft packing,mechanical seals were forced into cavityenvelope. Requires an API/CPI flushplant for optimal performance.B. Enlarged bore features increasedradial clearances over the standardbore. This chamber design enables opti-mal seal design. Restriction at bottom ofthe seal chamber limits fluid inter-change. Requires an API/CPI flush planfor best performance.C-D. Tapered bore features increasedradial clearances similar to the enlargedbore, except there is no restriction atthe bottom of the cavity and is open tothe impeller backside. Current designsinclude vanes or ribs to provide solidsand entrained gas handling. Flushing isoften not required as design promotescooler running seals by providingincreased circulation over faces.

4A

4B

4C

4D

Photo 3. Bearing housing configurations

Photo 4. Enhanced fabricated steelbaseplate

Page 150: Centrifugal Pumps Handbook

The Pump Handbook Series 151

TOMORROW'S CHEMICAL PROCESS PUMP

Having evolved for more than30 years, the ANSI/ASME chemicalprocess pump now offers usersimproved reliability, easier installa-tion and broader application flexibil-ity. Its development over the next 30years will surely produce furtherimprovements in these areas. Butinstead of 3 years meantime betweenplanned maintenance, the industrywill be driving towards 5+ years.Pump emissions will not be accept-able at 1000 ppm. Zero (0) ppm willbe the goal. It will not take 2 hours toalign a pump and motor to 0.002TIR. It will take only 15 minutes toalign to 0.0005 TIR. And these andother changes will surely take placein less than 30 years.

In fact, manufacturers and usersboth are demanding change now.The result is that current versions ofthe ANSI/ASME B73.1M andB73.2M are up for revision this year.The ANSI Pump Committee has alsodeveloped two new specifications. Inaddition, a new group called Process

Industry Practices (PIP) has devel-oped two new specifications thatsupplement ANSI B73 specificationswith additional requirements com-monly specified in the industry (bothhorizontal and vertical types).

1. POTENTIAL REVISIONS TOANSI/ASME B73 SPECIFICATIONS

Changes are under way in theANSI/ASME B73 Pump Committee.The basic B73.1M HorizontalProcess Pump Specification will berevised and probably issued in 1997.Areas being addressed to improvepump reliability are:

a. Nozzle loading b. Seal cavity dimensionsc. Auxiliary connections to glands

and seal cavities d. Baseplatese. Additional pump sizesf. Hydraulic Institute Class A

performance criteriag. Allowable operating range

New Specifications. The ANSI/ASME Pump Committee recentlyissued a new B73.5M Specification

addressing nonmetallic pump designs.In addition, a Canned Motor/MagneticDrive Specification B73.3M is expect-ed to be approved in 1996.

2. NEW PIP STANDARDThe new specifications covering

typical B73 pumps are the PIP(Process Industry Practices) RESP73Hand RESP73V. Engineering contrac-tors and pump users have formed aMachinery Function Team whosesole task is to develop a set of stan-dards that will eliminate variations inchemical process pumps manufac-tured to multiple user and contractorpump specifications. This team andits work will greatly minimize theproblems associated with multiplespecifications. Lower engineeringcosts and enhanced pump reliabilitywill be the benefits.

The PIP RESP73H (HorizontalChemical Process Pumps ANSIB73.1M Type) and RESP73V (Verti-cal Chemical Process Pump ANSIB73.2M Type) cover the same pumpdesign areas as the original ANSI specification, but they also address:

Pumpage Pumping Conditions Installation Considerations Materials of Construction

Corrosives Solids Hazardous Capacity Temperature Non-Metallic Metallic

Non (Noxious FRPAbras., Explosive Limited No or

Pump Fibrous Volatile Low High High Floor ANSI Align. PFA Poly- Alloy HighType Moderate Severe Stringy Abrasive Toxic) Flow Cap. Press. Cryogenic 0-500F. Sumps Space Dim. Req’d Teflon Tefzel* Propylene Iron Steel Alloys

VerticalSump

FRP Vert.Sump

InlineProcess

HorizontalProcess

TeflonLined

FRPProcess

Self-Priming

LowFlow

Non-Clog

HorizontalSeallessProcess

Non-MetallicSeallessProcess

Figure 5. Process pump selection guide

Page 151: Centrifugal Pumps Handbook

152 The Pump Handbook Series

a. Solid shaftsb. Shaft deflection L3/D4

c. Shaft sealing design responsibilityd. Bearing lubrication e. Preparation for shipmentf. Couplingsg. Baseplate designh. Hydraulic performance

acceptance criteria

The PIP team has the same goalsas the ANSI B73 pump team, but it isdriving standardization both to theproject and local levels. ■

Charles Cappellino is an engineeringproject manager for the Industrial Prod-ucts Group of Goulds Pumps Inc. He is aprofessional engineer with a Bachelor ofScience in mechanical engineering fromClarkson University and has beeninvolved in centrifugal pump analysisand design for 15 years.

Richard Blong is a product managerfor the Industrial Products Group ofGoulds Pumps Inc. He has a Bachelor ofScience in chemical engineering from theUniversity of Buffalo and has beeninvolved in centrifugal pump applicationsfor 10 years.

Page 152: Centrifugal Pumps Handbook

The Pump Handbook Series 153

ANSI Upgrades Require MoreThan TechnologyTotal program commitment is key to ANSI pump upgrade success.By Joseph Dolniak

Merely upgradingtechnology is notenough to increasepump reliability.A technologicalupgrade is just oneof many factorsthat must be add-

ressed to achieve maximum relia-bility. Reilly Industries has beenincreasing the reliability of its largeANSI pump population since 1990.At year end 1995 the company'stotal pump repairs were at their low-est levels since 1988. Meanwhile,plant production has tripled due toexpansions and improvements in effi-ciency.

We shall examine the proce-dures Reilly Industries instituted tobring its pumps into compliancewith the new ANSI standards. Spe-cific areas covered include newpump installations and orders,inventory, converting old pumps tonew standards and future plans.Sealless ANSI pumps are not cov-ered because there are none on site.Testing is currently under way todetermine if certain sealless brandswill be accepted into the plant.

A specialty chemical manufac-turer located in Indianapolis, ReillyIndustries is about to celebrate itscentennial under the leadership ofTom Reilly, Jr., the founder's grand-son. The company has grown overthe last century to where it currentlymanufactures more than 100 inter-mediate and specialty chemicals fora worldwide market.

Just over six years ago a pumpimprovement program was initiatedto increase the reliability of the com-pany's more than 800 pumps. Inimplementing pump upgrade pro-jects, certain steps must be followed

to obtain maximum benefit. Thesteps discussed below can be fol-lowed to complete any project suc-cessfully. They go by many namesbut most often are referred to asgood engineering practices.

GOOD ENGINEERING PRACTICESFollowing are some of the quality

steps or good engineering practicesfollowed in Reilly's pump reliabilityupgrade project.

1. Know the current situation2. Analyze the current situa-

tion3. Formulate a plan4. Initiate a trial5. Set up standard and use6. Train and communicate to

work force and others7. Maintain data8. Continue making refine-

ments9. Analyze pump technology

10. Phase in and phase out11. Stick to the plan12. Redo poor installationsSteps 1 - 8 are good engineering

practices overall. Steps 9 - 12 refermore specifically to the pump up-grading project.

For positive results in upgradingpumps for improved reliability, thepump, mechanical seal, gland,pump base, pump pad and immedi-ate pump piping all must beaddressed. The pumpage and flowrate conditions also must be compat-ible with the type of pump used.

When this project was started in1990, Reilly Industries had beenusing a computerized maintenancesystem for about 5 years. Thishelped greatly because the mainte-nance repair and cost historyalready was on file. The data indi-

cated that pump failures wereincreasing at an unacceptable rate.Production levels also were increas-ing, so there was an urgent need toimprove pump reliability. The his-torical data were only a portion ofthe information that needed to beanalyzed to implement a goodupgrade program. Visual inspectionsof the pumps installed at that timerevealed other factors that needed tobe addressed. The most obvious wasa lack of proper grout. Repairinspections also showed that manypump sites had excessive pipestrain. Indicator reverse, or laseralignment, was almost never doneon the ANSI pumps. These deficien-cies could be handled in-housethrough better maintenance prac-tices and training. Other difficultiescould not. The most important fac-tor that could not be controlled in-house was that mechanical sealswere operating in stuffing boxesdesigned for packing. But a newlyapproved standard known as ASMEB73.1M-1991 changed ANSI pumphistory.

ASME B73.1M-1991ASME B73.1M-1991 is the

"Specification for Horizontal EndSuction Centrifugal Pumps forChemical Process." One importantitem addressed in this revision wasan increase in pump base sizes toadd rigidity. Another was additionalmotor protection. Perhaps evenmore important, however, was thenew designation of a seal chamberversus the old stuffing box. Theintroduction of the seal chamberallowed the mechanical seal manu-facturers freedom to create newtechnology and mechanical sealdesigns. The gas barrier seals that

CENTRIFUGAL PUMPS

HANDBOOK

Page 153: Centrifugal Pumps Handbook

formulated to address pump reliabil-ity. Effort would be directed mainlyat ANSI pumps because of theirgreater population and the fact thatmany improvements made on themwould apply to other types of pumpsalso. The upgrade plan involved atwo-step approach. The first stepwas to ensure that no more pumpswere misinstalled or improperlyordered. This would help reducemaintenance problems from thestart. Reilly engineers began by writ-ing standards for proper pumpinstallation procedures. Covered inthe standards were such details aspump base dimensions, heights,depths, hold-down bolt designs,pump spacing from one another,pipe strain, grout and alignment.Before the standards wereapproved, they were tested on fourpumps that showed normal repairrates for the plant at that time. Afterthese pumps were reinstalledaccording to the proposed standard,they ran much quieter andsmoother.

Looking at the repair frequencyand costs for the three years beforethe reinstallation and after, totalrepair costs dropped 94%, from$46,470 to $2,908, and total repairsdropped 69%, from 49 repairs to 15in the same time period. Thesepumps were still fitted with the stan-dard stuffing box because the sealchamber was not yet on the market.

Shortly after the pumps were rein-stalled, however, the standard wasaccepted as a site engineering stan-dard for ANSI pump installation.

At the same time, new standardswere written for future ANSI pumporders. These standards incorporat-ed into all new pump orders many ofthe improvements specified in ASMEB73.1M-1991. Included were require-ments that all new pumps have 4ºtaper seal chambers, drain and dis-charge taps on the pump casing,labyrinth bearing isolators, bull's-eyeoil level indicators and bearing housingexpansion chambers. On pump-baseassemblies the base would conformto ASME B73.1-1991 di-mensions,have a grout hole centered on thebase (4" preferred), and have motorjacking bolts for alignment purposes.Some of these upgrades are shown inPhoto 1. This would give us a headstart as we would soon begin analignment and pump installation pro-gram. The jacking bolts and groutholes saved maintenance time whenthese programs were under way.

The accomplishments up to thistime took more than a year toachieve. They helped prevent misin-stallation of ANSI pumps and elimi-nate the process of ordering pumpswith old technology. This markedthe beginning of the next phase ofthe reliability program, but in actu-ality it would have little noticeableeffect for some time as the new

are now on the market are a resultof this. Specifics not addressed inthis revision were the standardiza-tion of the gland bolt circle, shaftsize at the mechanical seal, and seal-ing surface diameter at thegland/seal chamber interface andgland piloting area. Standardizingthese areas would allow the enduser even greater freedom to reduceme-chanical seal and gland invento-ry while maintaining competitivebidding from various mechanicalseal vendors.

With this revision, three basicseal chambers could be used. Theyincluded the 4º taper bore, the largebore with a throat restriction andthe large bore with no throat restric-tion. All three designs are beneficialto mechanical seals and can beswapped for the old stuffing box. Ineach the gland has to be replacedbecause of the larger static sealingsurface and bolt circle diameters.

PUMP RELIABILITY UPGRADE PLANWith the pump repair history,

installation analysis, repair analysisand the new ASME B73.1M-1991standard in place, a plan could be

154 The Pump Handbook Series

Photo 1. Recently received group two sizepump incorporates motor jacking bolts,grout hole, bearing isolators, bull's eyeoil indicator and bearing housing expan-sion chamber as standard features.

Photo 2. An upgraded second generation pump receives backup as part of the reliabilityimprovement program.

Page 154: Centrifugal Pumps Handbook

The Pump Handbook Series 155

pumps would be installed only asrequired through expansion or attri-tion. The next phase of the upgradehad more immediate impactbecause the ANSI pumps wereupgraded as they were worked on.This step addressed standardizingand consolidating pumps and pumpinventory while adding the new-technology parts to stores. This wasalso a two-step process. Step onewas the reduction in numbers ofANSI pump brands and the consoli-dation of mechanical seals. The sec-ond step consisted of phasing in newparts while the old parts werephased out of stores.

PUMP CONSOLIDATIONAt least nine specific brands of

ANSI pumps had been in use on site.The goal was to reduce that number tothree. Some of the factors consideredin selecting what brands to retainwere past reliability, maintenanceshop familiarity, local distributor pro-fessionalism, location of the OEM andthe size of the current plant popula-tion in specific makes. When selec-tion was finished, seven of the ninebrands were eliminated, and a newbrand was added. The standards listreflecting our preferred vendors wasupdated, and Reilly is now testing afourth brand of ANSI pump for possi-ble addition to the list. As for themechanical seals, we are about 95%committed to one manufacturer. Thissimplifies the process of consolidat-ing our mechanical seal inventory.

The time-consuming pumpbrand reduction project is just endingafter more than four years of phas-ing pump parts out of inventory.Caution should be taken whenadding additional brands of pumpsto a plant site once this point isreached because the amount ofwork needed to add a brand isalmost as time-consuming as it is toeliminate one. Also, the systemmust be allowed a break-in periodto determine how well it is work-ing. Continual changes do not allowthis to happen. Also, additionalshop training is needed. That iswhy the preliminary work of deter-mining what pumps will best suitthe plant is so important. After

finalizing the brands of pumpsaccepted on the vendor list, workbegan to phase in the acceptedbrands and eliminate brands thatwould no longer be used.

The company was careful notto be wasteful in eliminatingbrands. Repair parts in inventorythat were associated with the pumpbrands to be deleted were classifiedas POR (purchase on request). Thiswould cause no parts to be orderedwhen the reorder point came up,but it also would not delete the partfrom stores. In so doing, most ofthe repair parts currently in inven-tory for the brands of pumps to bedeleted could be used. If one ortwo minor parts were needed tocomplete a repair and the repairparts were not on hand, giving thestores clerk a work order numberwould allow the clerk to order thePOR part. When most of the partswere used up, the part code wouldbe changed to DELETE, and any re-maining parts would be pulled fromthe shelves. Thus, most of the partscould be used while eliminatingunwanted inventory. When no partsor not enough major parts remained,a new pump of an accepted brandwas ordered to replace the pumpthat was being repaired. The newpump incorporated all of theupgrades we required when itarrived in the plant.

Next month in Part II of thisarticle we will assess how ReillyIndustries conformed to the newupgrade plan. We will discuss con-solidation and parts inventorychanges, show how mechanicalseals were upgraded, and revealsome of the training procedures thathave been instrumental in helpingthis project succeed. We will endwith a look at future plans. ■

CONFORMANCE TO PLANIt would have been very easy at

the conformance stage to deviatefrom the plan because there weretimes when the up-front dollaramount involved in purchasing apart being deleted was less than thecost of upgrading the pump technol-ogy and purchasing a new pump.Economizing up front, however,

would have compromised the planand been costly in terms of overalllife cycle cost. This process is prov-ing profitable for the company asour overall pump repairs havedeclined as production has steadilyrisen, as shown in Graphs 1 and 2.Total repair costs for all pumps onsite appear to be going up slightly.This was expected because webegan requiring our work force todo more thorough repairs, whichtake more time. Maintenance work-ers now replace parts that were notnormally replaced due to lack ofproper inspection. More parts arenow being found to be out of specifi-cation when checked with dial indi-cators. And some auxiliary partsnow cost more because superiormaterials are being used. Yet if thepump repair costs were correctedfor inflation, the costs would bealmost level. Also, the total numberof pumps has increased. The in-house maintenance system showed792 pumps at the end of 1989 and837 pumps at the end of 1995.Achieving near level repair costswhile doing more thorough repairswith superior parts is attainablebecause we have increased ourMTBF (mean time between failure),and thus there are fewer repairs.

Another technique helpedmaintain costs while phasing outspecific pump brands. If a pumpbeing removed for an upgrade ortaken out of service was one of thebrands that was to be kept, and if itwas worth rebuilding, it was put ina specified area. If the pump filledthe requirements of a pump brandthat was to be eliminated, it some-times was used in lieu of purchasinga new pump. Several important con-siderations in doing this were thepumpage material and the repair fre-quency and the costs of the deletedpump brand being removed. Photo 1is an example of one of these pumpsbeing brought back into service andadding a back-up pump at this site,which before had no back-up pump.This helped determine if replacingold technology with old was appro-priate, or if new should be used, andit improved reliability while main-taining costs. To date, the equivalent

Page 155: Centrifugal Pumps Handbook

156 The Pump Handbook Series

of more than $50,000-worth of pumpparts has been removed from inven-tory records.

SEAL CONSOLIDATIONConsolidating the mechanical

seal parts was more difficult becausewe were phasing out standard stuff-ing boxes and phasing in 4º taperbore seal chambers. This requireddouble inventory of glands andgland gaskets for about two yearsbecause the gasket and bolt circle onthe seal chambers and their glandswere larger. There was no need toduplicate any mechanical seal parts.The mechanical seal stationaryinsert fit both the old glands usedwith the stuffing boxes and the newglands used with the seal chambers.Parts relating to the stuffing boxbackheads have recently been elimi-nated, and now only seal chambersand the new glands remain in inven-tory. There was another importantfactor in keeping the additionalinventory, now current inventory, toa minimum. In evaluating whatpump brands to keep, we alsolooked at the shaft diameter at themechanical seal, the seal chambergland bolt circle, and the gland gas-ket and pilot diameters. The threeANSI pump brands that were chosenhad the same dimensions for theabove items. This was important,and it is why, generally speaking,adding these items to ASME B73.1-1991 as part of the standard shouldbe beneficial to the end user. Bychoosing pumps with the samedimensions on the above items, wewere able to use only one gland to fitall three brands of pumps, per pumpgroup size designation.

Thus, for all of our group oneand group two size ANSI pumps,which are the majority of ANSIpumps in the plant, we have onlytwo glands: one for all group one sizepumps, and one for all group twosize pumps. The three types of com-ponent mechanical seals (one OEM)that we use per pump group size allfit the one gland. The glands comewith vent, flush and drain. If theyare not needed, they are simplyplugged. This has enabled us to elim-inate at least nine specific glands

from inventory. This strategy alsoworks for consolidating cartridgemechanical seals. The process tookmuch communication betweenmyself, the mechanical seal engi-neers and the pump engineers. Thesecond major benefit of consolidat-ing the glands besides the inventoryreduction is that there is no confu-sion to the work force on what glandto use on what pump, since there isonly one gland. This entire upgradeprocess is always evolving and beingupdated as needed. We are currently

consolidating three other glands intoone.

PARTS INVENTORY CHANGESOther changes made at this time

were the addition of the new tech-nology parts to inventory. Theseparts included solid shafts (versusthe sleeved shafts), bearing isolators(versus oil seals), bull's-eye oil indi-cator or column sight glasses (versusthe constant level), the new glandsand the 4º seal chambers (Photo 2).The old parts were earmarked so

Graph 1. Total plant pump repairs: REPPMP = repairs pump (one of a kind pumps),REPPMPP = repair positive displacement pumps, REPPMPV = repair vertical pumps,REPPMPC = repair centrifugal pumps, REPPMPTOT = total pump repairs

Graph 2. Total plant output using 1987 as the base quantity of one unit of output.

1200 —

1000 —

800 —

600 —

400 —

200 —

0 —

3.5 —

3 —

2.5 —

2 —

1.5 —

1 —

■—REPPMP■—REPPMPP■—REPPMPV■—REPPMPC■—REPPMPTOT

1987 1988 1989 1990 1992 1993 1994 1995

■■

■■ ■■

■■

■■

■■

■■

I I I I I I I I I I

■■

ESTIMATE

1987 1988 1989 1990 1991 1992 1993 1994 1995 1996

Page 156: Centrifugal Pumps Handbook

The Pump Handbook Series 157

that they could be phased out ofinventory and deleted at the appro-priate time. This created a few minorproblems because some of the pumpOEMs also were modifying variousparts because of the new ASMEB73.1-1991 standard. It caused us toadd parts that were being changed,and this created minor problems withassociated parts. The confusion wasminimized, however, by keepingcommunications open among the enduser, vendors and OEMs. As men-tioned earlier, this phase of the pro-ject is currently drawing to a close.

Because Reilly has a large num-ber of pumps installed that are twogenerations old, and were designedwhen mechanical packing was thenorm, they have inherent reliabilitydeficiencies when fitted withmechanical seals. These pumps havelong thin shafts and stuffing boxes.Due to their age and generation, thepump OEM will not be manufactur-ing upgraded parts such as sealchambers for these pumps. Becausethey represent a large part of theplant pump population, however, itwas important to improve their relia-bility also. Two primary points wereaddressed: seal environment andshaft stability. After conferring withthe OEM, it was agreed that thestuffing box could be bored. A newdiameter for the stuffing box wasdetermined, and the stuffing boxwas bored all the way through, giv-ing us what we call a modified largebore with no throat restriction. Ourlocal vendor for this pump brandsees to it that these parts now comeinto stores already bored throughwhen they are ordered from theOEM. To increase the shaft stiffness,we opted for a solid sleeveless shaft.

This type of part modification isalso being applied to some of ourlarger group two and group threesize pumps, which represent aminority of the plant pump popula-tion. Because of their limited num-bers, we stock only commonly usedrepair parts (shafts, bearings,mechanical seals). Large-dollar partssuch as casings and seal chambersare listed as POR. By boring out thestuffing box and keeping the stan-dard gland, we are able to receive

some of the reliability benefits of thetaper, or large bore seal chamber,while holding down repair costs.When there is need for more com-plete repair, these pumps also willbe converted to updated technolo-gy.

MECHANICAL SEAL UPGRADESUpgrading the mechanical seals

was another important process forus. It went hand in hand with con-solidating the mechanical seals. Spe-cific failures were noted on one typeof elastomer. Other failures specificto one plant were noticed on one ofthe seal hard faces. Also, with morestringent regulations approaching,we felt it would be beneficial toupgrade the seal/elastomer combina-tions used throughout the plant.Because our ANSI pumps generallywere not pumping "easy" products,we decided that it would be best touse premium hardfaces and elas-tomers for all mechanical seals onANSI pumps. Due to our consolida-tions, this affects only eight specificseals. The reasoning was that ouraverage pump repair plantwide(ANSI and non-ANSI) was about$1,000 per repair. This covers anyaction ranging from no repair orminor repair to installing a newpump. Significant failures werenoticed on encapsulated o-rings.Converting to an elastomer such asKelrez added several hundred dol-lars to the cost of the mechanicalseal. Still, if repairs were reduced,costs would decrease in the long runbecause there would be fewer fail-ures (Graph 1). The new hardfaceused to replace the hardface experi-encing problems in one specificplant would also work throughoutthe entire Indianapolis site. There-fore, this conversion was complet-ed as well.

Through failure analysis, run-ning dry was determined to be oneof the major factors contributing topremature seal failure, even afterpump upgrades were completed. Toretain the gains discussed here, threeop-tions were identified to addressthis problem. The first is to fit pumpmotors that can be turned off whilethe pumps are in service with power

monitors. This pertains primarily totransfer pumps. Use of properly cal-ibrated power monitors has all buteliminated seal failures at specificpump sites within our plant. Thesesites also have been correctlyupgraded, which eliminates earlierroot causes of pump failures beforethe power monitors were installed.The second option involves replac-ing the 4º taper seal chamber with arestricted throat large bore sealchamber. Swapping seal chambers isconsidered because of the pumpageproperties in certain pumps. Thereare signs of running dry, whichcould actually be entrained air, or aphase change of the pumpage at theseal faces. The restricted throat largebore seal chamber with various flushplans allows us to obtain a higherpressure in the seal chamber andthus reduce or eliminate this prob-lem. The third option to preventpumps from running dry pertains toareas where the pump cannot beturned off automatically because ofproduction needs. For some of theseareas, we are installing gas barrierseals designed for the pumpingrequirement. Until recently the gasbarrier seals required that a sealchamber be present, as they wouldnot fit in the common stuffing boxes.This situation is changing and gasbarrier seals are becoming availablefor the use in stuffing boxes (seePumps and Systems, February 1996,pages 8 and 9). As of yet, we haveordered and received only one gasbarrier seal, and we are waiting toinstall it when the current pumpcomes out for repair.

TRAININGOne aspect not yet mentioned is

training for the maintenance workforce. Workers were generally con-cerned about all of the changes. Newbrands of pumps were being used.Different types of technology wererepresented. More was expected ofworkers. All of these were realissues that had to be addressed.Although one-on-one discussionsand shop meetings were held, theonly way to ensure full communica-tion of all of the changes wasthrough formal training for the

Page 157: Centrifugal Pumps Handbook

158 The Pump Handbook Series

maintenance work force. Formaltraining assures us that all of ourworkers are up to date on new partsand procedures. Specific trainingareas have included ANSI pumprebuilding, mechanical seal installa-tion and indicator reverse alignment.

Training in ANSI pump rebuild-ing covered the specific dial indica-tor checks that must be performed toensure that the pump componentsare in the proper operating specifica-tion. Mechanical seal installationtraining noted the proper techniquesto install both component and car-tridge mechanical seals by the print,and by the stack method, and testingthe seals on the bench beforeinstalling the pump in the plant.Indicator reverse alignment trainingcovers how to align pumps properlyand overcome common alignmentproblems such as soft foot, bolt bind-ing and shim pack requirements.

OEM AND VENDOR TRAININGTraining the work force was

only part of the project. Training thevendors and OEMs and maintaininggood communication among themwere critical parts of the trainingprogram. This meant keeping themup to date on our new pump require-ments, allowing them to read andunderstand our plant engineeringstandards relating to the pumps, andensuring that they held enoughpump part inventory to cover ournormal use. This was especiallyimportant in the beginning stagesbecause being out of a specific newtechnology part needed for a repair,especially on a critical pump, wouldadd fuel to the fire of the nay-sayers.This was true more so when weimplemented the changes becausemany of the pump OEMs were notyet willing to stock taper and largebore seal chambers because there"was no demand" at the early stagesof their introduction to the market.My reply was that we were demand-ing them, so please stock them.Because these new parts haveproved to be important in increasingpump reliability, and there is indeeddemand for them, most vendors nowstock these parts.

When this program started, one

of the major components of upgrad-ing the ANSI pumps was the newseal chambers. Reilly first upgradedthe bulk of its ANSI pumps to thenew ANSI standard. We felt that thisalone would increase the pump relia-bility. When the general populationwas updated, more specific prob-lems would become clear and couldbe addressed. Looking at the threeoptions that were available, it wasdecided that the taper bore sealchamber would best fit the overallneeds of the company. Because ofthe many other factors that neededto be addressed to get the pumps,stores, and the vendors up to date onthe pump upgrades being imple-mented, only one seal chamber waschosen Adding two seal chambers tothe system would be detrimental tothe entire process, and it would justadd confusion until the project wasmore mature. Now that this generalpump upgrade is accomplished, asmentioned earlier, we are looking atcertain specific installations wherethe large bore seal chamber with therestricted throat might be of morebenefit. This is primarily because ofthe makeup of the product, and notbecause of gland erosion. We haveseen only one instance of gland ero-sion as a result of the cyclonic effectof solids and entrained air in a sealchamber, and the one instance wasin clean acid. We have also had ourmechanical seal OEM perform testson various flushing arrangementsusing the taper bore seal chamberand our own gland (group one sizepumps, 1800 rpm) with respect toparticulate and entrained air in theseal chamber, and we have hadinteresting results that we can direct-ly apply to field use. With conver-sion to the 4º taper bore sealchambers from stuffing boxes wellestablished, if we now want tochange from a taper bore to a largebore seal chamber, the gland, sealand other pump parts will fit. Weneed only change the type of sealchamber used for the specific pumpin question.

FUTURE PLANSOur ANSI pump upgrade pro-

gram has been operating for near-

ly three years now. But along withupgrading the pump technology,we also had to address the rootcauses of failures. Reilly has begunto eliminate pipe strain proactive-ly. Completely reinstalling prob-lem pump sites has been accom-plished with good results. Some ofthese sites incorporated bothepoxy bases and epoxy grout,which give superior chemicalresistance and vibration dampen-ing (Photo 3). Steps have been tak-en to improve our groutingtechniques, and we are completelyphasing in epoxy grout. Comparedto cement types of grout, this givessuperior vibration dampening,chemical resistance and adhesionwhen the pump base and pad com-ponents are properly prepared.Also, bearings with tighter toler-ances than we normally stock arebeing phased into stores.

Because of the long term results,there are now more requests fromupper management to apply thepump reliability and upgrading tech-niques further – to specific problempumps in each plant. This willinclude improving our vibrationanalysis program and reiterating theimportance of aligning every pumpthat is worked on, no matter whatthe size. As pump reliability contin-ues to improve, new "bad" pumpswill make the list of pumps to beaddressed as old ones are removedfrom the list because of improvedreliability. This is part of the contin-uing refinement. By analyzing pumpfailure data, improvements can con-tinue as long as there is a wellthought out improvement process.Looking at the repair graph, it mayappear that our reliability is improv-ing at a slow rate. But in looking atthe entire picture, our plant hasbeen undergoing extensive growththrough expansion and efficiencyimprovement. Our production hassteadily increased, our pump popu-lation is increasing, new productsare being brought to productionwhile production of older core prod-ucts are being reduced. Thesechanges have created new chal-lenges, and such conditions can havea dramatic effect on the pump and

Page 158: Centrifugal Pumps Handbook

The Pump Handbook Series 159

mechanical seal population of anyplant. When all these points are con-sidered, however, there is significantsatisfaction with the progressachieved in a pump reliability pro-gram of this type. ■

Joseph Dolniak has been themaintenance engineer at ReillyIndustries for more than 6 years andis involved in reliability improvementsof rotating equipment. In addition toproducing alignment, seal installationand pump rebuilding training videosfor Reilly, Joe has published severalarticles on ANSI pump reliability andhas lectured at the Pump UsersSymposium and the Pump Reliabilityand Maintenance Conference.

Photo 3. Upgraded pump site includesepoxy base and grout for improved reliability.

Page 159: Centrifugal Pumps Handbook

160 The Pump Handbook Series

Selecting Mag Drive PumpsMagnetic drive pumps offer irresistible force in sealless pumping.

By Robert C. Waterbury, Senior Editor

Industrial processes involvingtoxic, hazardous or environ-ment-threatening chemicalsoften employ the magnetic dri-ve pump (MDP) as a safe, seal-less solution. But even thoughMDPs offer a simple answerto a common need, certain

characteristics must be considered toselect and apply them cost-effective-ly.

Kaz Ooka, president and found-er of Ansimag, Inc., points out thatmagnetic drive pumps historicallydeveloped along two lines: metallicand non-metallic. The metallic de-signs traditionally were used inprocess or heavy-duty applications.But non-metallic pumps, once con-sidered only for light duty applica-tions, have moved up in power andsize due to development ofimproved rare earth materials suchas samarium-cobalt and neodymi-um-iron-boron.

Synchronous MDPs use rareearth magnets. Because they areaffected by high temperatures, theyoften require special cooling provi-sions for applications in excess of400ºF. Eddy current MDPs employa torque ring that is normally unaf-fected by temperatures found in hotoil heat transfer systems. They use arotating assembly sealed by a con-tainment shell. Power is transmittedby permanent magnets mechanical-ly coupled to the driver rather thanthrough motor windings.

Sealless MDPs prevent liquidleakage and eliminate common envi-ronmental concerns. They are alsoused to move liquids that crystallizeupon contact with air, and the sealflush liquids or gases they employhelp avoid contamination of processfluids. Yet they don't solve all seal-

related problems. Two issues thatMDPs still must address are mini-mum flow conditions and dry-run-ning. Minimum flow rate is greatlyaffected by radial or thrust load onthe bearings or pump shaft and thetemperature rise. Dry-running is themost common cause of failure inMDPs and results in thermal dam-age to the metallic containment shelland/or in mechanical or thermalshock to the bearings and shaft. Weshall discuss these in more detail.

DESIGN CONSIDERATIONSThe inside of a sealless magnet-

ic drive pump reveals a complexinternal flow system that is difficultto model. However, the internaldesign holds the key to cooling themagnet drive and lubricating thebearings effectively and to the safetransport of solids. Engineers atHMD Seal/Less Pumps in East Sus-

sex, England, which is affiliatedwith Sundstrand Kontro, identifyfive critical design elements:

1. the liquid end, comprisingpump casing and impeller

2. a magnetic drive includingan inner and outer magnet assemblyand the containment shroud

3. internal support bearings4. an internal feed system that

circulates among 1, 2 and 3 aboveand is needed to cool the magneticdrive, lubricate the bearings andtransport any solids in suspension

5. a power frame that compris-es the external bearings supportingthe outer magnet and the interfaceto the prime mover

INTERNAL FEED SYSTEMOf these, HMD considers the

internal support bearings and inter-nal feed system the most critical,and yet perhaps least discussed. The

CENTRIFUGAL PUMPS

HANDBOOK

Figure 1. Rotan MD series magnetic drive pump

ROTAN MD SERIES PUMP

Page 160: Centrifugal Pumps Handbook

The Pump Handbook Series 161

same effect as high flow and exces-sive pressure drop in causing vapor-ization.

INTERNAL BEARINGSMichael "Todd" Stevens, senior

maintenance engineer at HoechstCelanese Chemical in Houston, hasanalyzed bearing failures in magnet-ic drive pumps and offers somehelpful observations. He suggestslooking at the lubrication scheme forthe roller/ball bearings and the shaftseal that prevents lubricant fromentering the drive magnet section ofthe pump. When using wet sumplubrication, shaft failure will allowan oil level to build in the drive mag-net section, causing a heat buildup.Normally under these conditions thepump will begin making noise andvibrate. If the oil level buildup is notdetected, the temperature of the dri-ve magnet section will increase andeventually cause a roller/ball bear-ing failure.

Stevens also suggests selectingsilicon carbide rather than carbon asa main bearing material. It has bet-ter wear properties and will with-stand most thermal shock withoutfailure. Some suppliers even dia-mond coat silicon carbide journalbearings to withstand brief periodsof dry-running without bearing fail-ure.

Thrust bearings should be engi-

neered so that full-face contact isachieved between the bearing andthrust runner. Failure to do soresults in point loading, which candamage the bearing. AlthoughStevens recommends a fixed facebearing, he concedes that a floating-face spherical-seated thrust bearingcould be acceptable if there weresome way to secure the floating faceat all times. Also, open impellersimpose higher thrust loads thanclosed impellers. Thus, thrust loadsmust be either offset by balancingthe impeller or absorbed by thethrust bearings. Lower thrust loadsobviously mean increased bearingreliability.

INTERNAL HEAT GENERATIONBearings are not the only mech-

anisms that generate heat in a magdrive pump. In fact, Stevens pointsout that most heat is actually causedby eddy current losses between thedriven and drive magnets. With apermanent magnetic coupling, thedriven rotor turns at the same speedas the drive rotor. The two magnetsare separated by a containmentshell. Heat generation is thus a func-tion of pump rotation speed andcontainment shell construction.Containment shell materials such as316 stainless steel generate moreheat than Hastelloy C, Stevensnotes, and Hastelloy C generates

feed system removes heat generatedin the drive assembly by eddy cur-rent and viscous friction losses, andit lubricates the process lubricatedbearings that support the loadsexerted upon the rotor assembly.

Discharge to suction internalflow. To remove heat produced inthe drive assembly, liquid is takenfrom the discharge of the pump andreturned at a lower pressure pointwithin the pump. The simplest andperhaps lowest cost system takesprocess liquid from the pump dis-charge and recirculates it to thepump suction end. Typically,process liquid leaves the exit of theimpeller and returns to the magneticdrive through a hole in the rear cas-ing plate. With this system the liq-uid always returns to the suction ofthe pump at a higher temperaturethan the bulk suction temperature.Thus, it is necessary to ensure thatthe hotter temperature liquid doesnot affect the NPSHR of the pumpand vaporize as it returns to theimpeller eye. This condition is oftenoverlooked, according to HMD, ifsuppliers test using only water.Because water has a liquid with ahigh specific heat and gradual vaporpressure curve, test results usingwater alone may mask this potentialproblem.

Discharge to discharge inter-nal flow. An alternative is to takethe discharge liquid from the pumpand return it at a point of pressurehigher than suction. The internalfeed system is similar to the dis-charge to suction feed system,except that the liquid is directed to ahigh pressure area in the casing typ-ically behind the back shroud of theimpeller. There are two advantagesto this system. First, it eliminatesthe NPSHR problem. Second, pres-sure distribution within the drive isrelated to a high pressure area asopposed to suction and thereforereduces the possibility of vaporiza-tion within the drive. The main dis-advantage is that the supplier musttest to ensure that there is adequatepressure difference under all condi-tions to force sufficient flow throughthe drive. If not, low flow and exces-sive temperature rise will have the

Photo 1. Mag drive pump transfer of 93% sulfuric acid has operated nearly two yearswithout repair

PHOT

O CO

URTE

SY O

F AN

SIM

AG, I

NC.

Page 161: Centrifugal Pumps Handbook

162 The Pump Handbook Series

more heat than high performanceplastics. The temperature of the con-tainment shell directly between thetwo rotating magnets can easily riseabove 750ºF within 30 seconds ofthe onset of dry-running, and it caneventually reach nearly 1000ºF. Sointernal fluid flow between the dri-ven magnet rotor and the contain-ment shell is needed to remove theheat.

Dry-running heats the shaft. Ifcool liquid is introduced at this time,however, the shaft and bearingsmay fail due to thermal shock.Ceramic materials can minimizethese effects but offer widely vary-ing thermal shock limits. Ooka saysalumina ceramic can withstand onlya 200ºF thermal shock while sin-tered carbide offers resistance up to600ºF. Silicon carbide offers suchhigh thermal shock resistancebecause it has a very low coefficientof thermal expansion combinedwith a very high thermal conductiv-ity. This allows the material toequalize in temperature very quick-ly while exhibiting minimal thermalstrain.

MDP MONITORING/INSTRUMENTATIONSafe, reliable operation of MDPs

clearly depends not only on pumpselection and installation, but onmonitoring pump operating condi-tions. Following are some of themore common techniques andinstrumentation.

Temperature monitoring.One way to monitor pump conditionis to use a thermocouple or RTD(resistance thermal device). It can bepositioned to monitor the tempera-ture of the containment shell orplaced in a thermowell to indicatethe temperature of the fluid leavingthe shell. Either way, it monitors theheat produced by the eddy currentlosses in the magnetic coupling aswell as the bearing friction.The tem-perature can be used as an absoluteor a differential measurement. Usedas a differential temperature indica-tor, it is referenced to the pump suc-tion fluid temperature and directlymeasures heat input to the fluid bythe pump.

Flow protection. Minimumflow protection is normally providedby installing a flowmeter in the dis-charge line of each pump. Stevenspoints out that the minimum flow toprotect against is either the thermal orthe stable minimum flow, whicheveris greater. It is only reliable, however,if no more than one pump is operat-ing in a two-pump system. Other-wise, each pump must beindividually instrumented and pro-tected.

Low suction. Low suction ves-sel protection ensures that a pumpwill not run dry. According toStevens, it is used in tank loadingand offloading applications toensure that the pump will not rundry or suffer from inadequateNPSHR. If a tank must be emptiedfollowing unloading, then a mag dri-ve pump should not be used. Other-wise cavitation and subsequentfailure of the thrust bearings andpossibly the sleeve bearings couldresult.

Power monitoring. Technical-ly, a power monitor can help guardagainst low flow, high flow, magnet-ic decoupling and dry-running. Itmeasures the power consumed bythe motor and thus responds quicklyto load changes that could lead tomechanical damage. It obviouslyapplies only to pumps driven byelectric motors, however, andwhether it is sensitive enough to dis-tinguish between low flow and shut-off is questionable. In such cases aflowmeter or other device may berequired as backup.

Vibration monitoring. Sleevebearings typically run so smoothly(less than 0.1 in/sec overall) that peri-odic vibration monitoring has notproven useful in predicting failure. Ithas been successful, however, in pre-dicting the failure of the drive mag-net support bearings – normallyroller or ball bearings. This couldhelp ensure that the drive magnetdoes not contact the containmentshell in case of a support bearing fail-ure.

THE DECISION PROCESSThe Clean Air Act targets 179

liquids in setting limits for allowablechemical leakage into the air. Theprimary concern is for safety ofhumans and the environment. Andmagnetic drive pumps eliminate thedangers normally associated withseal leakage in mechanical pumps.But even though MDPs are pur-chased initially for safety reasons,many users are now specifying themfor reasons of improved reliability,extended service life and longermean time between maintenanceand repair. In the long run they mayprove more economical even thoughthe initial cost is higher. Once thedecision is made to purchase a mag-netic drive pump, however, manyquestions must still be answered aspart of the selection process. The fol-lowing selection guide developed byAnsimag can help users tailor solu-tions to meet their specific needs.

Dimensions and design. Doesthe manufacturer make a pump in

Photo 2. A mag drive pump transferring50% sodium hydroxide at a specialty chemical manufacturer in the southeast-ern U.S.

PHOT

O CO

URTE

SY O

F AN

SIM

AG, I

NC.

Page 162: Centrifugal Pumps Handbook

The Pump Handbook Series 163

the design that you require? Thereare numerous configurations andstandards including ANSI, ISO, API,DIN, etc.

Solution: Determine the designthat you need and consider onlythose manufacturers that build thattype of pump.

Vapor pressure of the liquid.Metal magnet drive pumps add heatto the process fluid due to losses inthe magnetic coupling. The typicalmagnetic coupling is anywhere from70% to 80% efficient. This inefficien-cy is translated into BTUs that enterthe liquid as temperature. As muchas 20ºF can be added to the liquidthat is lubricating the bushings. Ifthe liquid vaporizes, the bushingswill be starved for lubrication, andthe pump will fail.

Solution: Ask the manufacturerto run a heat balance calculation todetermine if the vapor pressure of theliquid will ever exceed the local pres-sure in the pump. If it does, it is thewrong pump. Also, consider a non-metallic magnet drive pump withzero losses in the magnetic coupling.This will eliminate the possibility ofvaporizing the liquid.

Solids. Because the bushingsare lubricated by the process fluid,a "clean" liquid is required. Somepumps will handle more solidsthan others.

Solution: Ask manufacturers tostate maximum limits and give refer-ences of applications handling simi-lar solids content. A general rule ofthumb is 5% by weight and 150microns maximum. Flush systems areavailable from some manufacturers toincrease solids handling capacity.

Hydraulic capacity. Somemanufacturers have more capabilitywith regard to head and flow thanothers because they offer more mod-els.

Solution: Look at what modelsthe manufacturer has available andready to ship (not just planned asfuture products but currently avail-able). Even if the manufacturer haswhat you need currently, consideralso that you may want to add larger

units at a later date. If he does nothave them available, you lose com-monality of parts and continuity ofdesign.

Temperature. Magnet drivepumps have a variety of temperaturecapabilities.

Solution: Determine not onlywhat temperature you will be oper-ating at, but also what maximumtemperature the pump could see dueto excursions or future design revi-sions. Also, consider the effects ofsteam cleaning or heat tracing ifyou plan on either of these. Look atboth the magnet capability as well asthe material capability of the pumpwith regard to maximum tempera-tures.

Simplicity and ruggedness.These two items are critical now thatmaintenance staffs are slimmer. Theless time spent on the pump the bet-ter – ruggedness of design is key.When maintenance is required, thesimpler the better since the timespent on repair should be minimizedand the risk of making an errorshould be reduced.

Solution: Ask the distributor ormanufacturer to demonstrate thepump to determine if it is indeedsimple and rugged. Viewing thepump in operation is critical becauseevery manufacturer claims to have asimple and rugged design.

CAUTION: RISK AHEADAs Stevens says, process engi-

neering people design systems fornormal operation and project engi-neers then use these flow require-ments to purchase pumps. This is anormal method of sizing and pur-chasing a pump, but it is not alwayssuccessful in purchasing and sizing amag drive pump. An MDP is some-what more sensitive to changes inpumping conditions than perhaps anANSI design pump. Startup proce-dures, for example, do not alwayscall for operation using the same flu-id, pressure, temperature, specificgravity and viscosity indicated onthe data sheet. Furthermore, pumpsmay be used for more than one oper-ation, or they may be required to

pump at widely differing flow ratesduring unit startup, operation andshutdown cycles. The greatest risksare always posed by conditions thatfall outside the normal range of oper-ation.

The pump system is designed tooperate at the Best Efficiency Point(BEP). However, real world condi-tions demand more from a pumpthan a single BEP. A pump may beused to transfer fluid from tower totower before unit startup to achievea normal tower operating level. Sim-ilarly, it may be used to clear theunit in case of a unit trip, or it maybe used as a spare for a completelydifferent service via a jumper line.These scenarios must be postulatedand the implications explored beforeinstalling a magnetic drive pump.Properly done, this exercise will pro-vide an operating window of mini-mum/maximum values to beconsidered in the selection process.

Improper application is perhapsthe most frequent cause of failure.Calculations of available versusrequired NPSH must be extremelyaccurate and compatible; otherwise,cavitation and pump failure will fol-low quickly. In addition, obviouspractical considerations such asdirection of motor rotation must notbe overlooked. In a recent pumpstartup operation, three of nine ini-tial failures were due to incorrectmotor rotation.

SELECTED PRODUCTS ANDAPPLICATIONS

Ansimag. Different versions ofthe Ansimag K1516 mag drive pumpare being used to move an extensivelist of hazardous, corrosive and toxicchemicals. Applications noted mostfrequently in a new Ansimag casehistory publication involve suchchemicals as hydrochloric acid, sul-furic acid, sodium hydroxide andsodium hypochlorite. The main rea-sons users give for switching to thesepumps include zero leakage require-ments, safety, elimination of sealproblems and increased systemuptime. The users include specialtychemical and petrochemical compa-nies, pulp and paper processors,food and pharmaceutical companies,

Page 163: Centrifugal Pumps Handbook

164 The Pump Handbook Series

and steel, plastics and electronicsmanufacturers. In its list of userapplications, Ansimag records userchemical concentrations rangingfrom 1 -100%, flows from 1-500 gpmand TDH in feet from 5-250. Processtemperatures generally range from -94º to 250ºF.

Kontro/HMD. Kontro andHMD Seal/Less pumps are availablein a wide range of capabilitiesdesigned for specific target applica-tions. The A-Range mag drive pumpsfeature capacities to 2000 gpm,heads to 700 ft TDH, temperaturesto 400ºF and system pressures fromfull vacuum to 275 psig. Applica-tions include toxic or hazardous liq-uids, high temperature vacuumdistillations and liquids that areexpensive or require controlled puri-ty. The H-Range pumps offer capaci-ties to 5000 gpm, heads to 700 ftTDH, temperatures to 750ºF andsystem pressures from full vacuumto 300 psig. Applications includeheat transfer fluids, molten solidsand high temperature vacuum distil-lations. The API pumps range to5000 gpm capacity, heads to 700 ftTDH, temperatures to 750ºF andpressures to 580 psig. Applicationsare refinery and petrochemical ser-vices. The HSP series fills highsystem pressure requirements in-cluding: capacities to 2000 gpm,heads to 350 ft TDH, temperaturesto 750ºF and system pressures to5000 psig. These pumps are used innuclear, high pressure densitometerand pipeline detection systems.Finally, the self priming SP pumpsare designed for truck and tank caroffloading of hazardous chemicals.They accommodate capacities to 350gpm, heads to 300 ft TDH, tempera-tures to 400ºF, pressures from fullvacuum to 150 psig and suction liftto 15 ft.

Micropump. Micropump offersprecision fluid pumps and systems.Its Integral Series line is used inhemodialysis, chemical dosing,dispensing and filling, waterpurification, ink jet printing andlaser/electronics cooling. Its distinc-tive features include motor, variablespeed control via external signal, low

power consumption, pressures to100 psi and flow rates to 5.8 gpm.The pump head, brushless dc motorand electronic controller are inte-grated into a single compact unitwith no mechanical seals, packing orleakage. The drive system is distin-guished from conventional magneticcouplings by sealing the rotorinside the pump and driving itdirectly by the motor stator. Theelectronic controller, an integralpart of the motor drive, acceptsseparate 0-5 Vdc or 4-20 mA signalsused to adjust speed control.

DESMI/Rotan. Mag drive seal-less pumps in the MD series offercapacities to 225 gpm, speed to 1750rpm, differential pressure to 250 psi,temperature range to 500ºF, suctionlift to 15" Hg vacuum while primingand 25" Hg while pumping. MDpumps are recognized for their inte-gral pump cooling system, dynamicaxial balancing feature that reducesenergy consumption and increasesMTBM, a thrust control system thatmaintains correct running clear-ances and reversible pumping capa-bility through changes in motorrotation. A patented system circu-lates the pumpage around the magnet-ic coupling for cooling. A specialshaping at the rear of the rotor usesthe hydraulic pressure itself to bal-ance the liquid pressure dynamicallyon the rotor.

Maag Pump. MPS pumps fromMaag are known for their use invery high pressure applications.Their operating conditions featuretemperature ranges to 300ºC, suc-tion pressure to 16 bar vacuum, dis-charge pressure to 66 bar maximum,differential pressure to 50 bar maxi-mum and viscosity to 1000 m Pa.They are available with either singleor double containment shells.

Price Pump. The model CD100MD from Price pump offersflows to 70 gpm and heads to 95 ftTDH. Incorporating 316 stainlesssteel or higher alloy materials, itsstandard configuration withstandspressures to 300 psig and tempera-tures to 350ºF. Factory engineer-

ing is available for higher tempera-ture and pressure ratings. Threemagnet strengths are available forvaried load conditions, and insru-mentation options include powermeter, temperature probe, andvibration, temperature and pressureswitches.

Roth Pump. Magnetic drivepumps from Roth offer the advan-tages of regenerative turbine pumpsin addition to a low NPSH feature. Afloating, self-centered impeller isable to produce any level of differen-tial pressure from 50 to 500 ft TDH.Maximum pressure with 316 stain-less steel (standard) is 230 psi, or upto 360 psi with optional Hastelloy Cmaterial. A power factor sensor indi-cates both high and low load andreacts to upsets caused by blockedvalves or vapor-bound conditions.

Dean Pump. The M300 con-forms to the dimensional specifica-tions of ANSI B73.1 and features aone-piece hydroformed containmentshell. The seal for this shell is theonly o-ring in the pump, and option-al flow paths are offered to meet spe-cial situations. The RM5000 is aheavy-duty process pump with acenterline mounted refinery typepump wet end. This design offershigher head and capacity rangesalong with optional flow paths. Thecontainment shell is gasket sealedwith no o-rings. The RMA5000 is ahigh temperature variation that canbe air cooled to pumpage tempera-tures of 750ºF. It features the exter-nal-external flush system and usesno o-rings.

Klaus Union. A wide variety ofmag drive pumps conforming toANSI, API, and DIN specifications isavailable from Klaus Union. Theyaccommodate heads up to 575 ft,temperatures from -300°F to 840°Fand design pressures to 5800 psig.Multistage high pressure pumpsoffer total delivery heads to 3300 ft.A patented double isolation shell isnoteworthy among its offerings inaddition to a pressure switch for con-tinuos monitoring. Standard prod-ucts offer designs geared to toxic

Page 164: Centrifugal Pumps Handbook

The Pump Handbook Series 165

duty, high temperature, high pres-sure and slurry applications. Productsare available in single and multistagehorizontal, vertical and even screwpump configurations. Additionalheating and cooling is provided byjackets or coils, and forged casingsand special isolation shells are used inhigh pressure designs.

PUTTING IT ALL TOGETHERIn Michael "Todd" Stevens'

experience "...every mag drive pumpfailure has been the fault of somesystem design or operation upset.The mag drive pumps, when operat-ed properly, have been very reli-able." So how can one ensure thebest continued operation of magnet-ic drive pumps?

The answer is not necessarilysimple. First, Rob Plummer of DeanPumps suggests that all mechanicalseal problems really need to beresolved before even considering thepurchase of a mag drive pump. Hisreasoning is that most of the materi-als used in mag drive pumps are thesame used in mechanical seals —just in a different manner. Further-more, a mechanical seal is tolerant ofcertain design features such as an

elbow on the suction of an end suc-tion pump. A mag drive pump instal-lation, on the other hand, requires astraight pipe with smooth flow. Andof course there are the dry-runningand eddy current heat problems thatwe have discussed.

Having considered these issuesand still electing to install a mag dri-ve pump, one must closely analyzeall possible operating conditions(including startup and shutdown)and pass that information on to thesupplier. This involves specifyingsuch characteristics as type of chemi-cal(s), specific gravity, viscosity, spe-cific heat, vapor pressure/temperaturerise, percentage of solids and dis-solved gases. Describe all possibleprocess operating conditions. Pro-tect the pump by using at least atemperature sensor to indicate thefluid heat as it leaves the contain-ment shell and a power monitor toindicate internal operating condi-tions. Enlist the assistance of main-tenance personnel in system design,pump selection, startup, trainingand general operation. And be sureto check obvious details such as thedirection of motor rotation! ■

Page 165: Centrifugal Pumps Handbook

166 The Pump Handbook Series

Operation Protection for Mag Drive Pumps Learn how to avoid dry or semi-dry running conditions, which can lead to damage.By Kaz Ooka and Manfred Klein

Magnetic drive centrifugalpumps are products ofan evolving technologyutilizing new materials,stronger magnets andnew concepts. Begin-

ning shortly after World War II, themag-drive concept developed alongtwo paths – namely, metallic andnon-metallic pumps. Metallicdesigns have been utilized primarilyas process or heavy duty pumps,notably in Europe. In earlier years,non-metallic mag drives were usual-ly considered applicable in lightduty situations only – fish tankpumps drawing 100 watts or less,for example. With the developmentof rare earth magnet materials suchas samarium-cobalt and neodymi-um-iron-boron, however, the sizeand power of non-metallic designshave been greatly improved over thelast ten years.

This rapid increase in magnetstrength has allowed for a corre-sponding reduction in the size andweight of the magnetic coupling.Photo 1 shows an industrial non-metallic mag-drive pump. Thismachine is clearly a great improve-ment over the original fish tankpump.

As more magnetic drive pumpshave been applied in the processindustries to solve increasingly com-plex problems, some confusion hasdeveloped among users accustomedto sealed centrifugal pumps such asthose described in the ANSI/ASMEB73.1 standard: Specification forHorizontal End Suction CentrifugalPumps for Chemical Process. Manyusers have believed that due to itssealless cocnstruction the magneticdrive pump can solve all seal related

problems. Sealless pumps are idealfor preventing liquid leakage andmitigating the associated environ-mental concerns. They also workwell in the pumping of liquids thatcrystallize upon contact with air,and they avoid contamination ofprocess fluids by seal flush liquidsor gases. And they are excellent forpumping corrosive liquids. Howev-er, several issues need to beaddressed, the most important ofwhich are minimum flow conditionsand dry-running. While these issuesare relevant to traditional sealedpumps as well, this article will con-centrate on their effects on mag-dri-ve pumps.

MINIMUM FLOWTwo factors determine mini-

mum flow rate: radial or thrust loadon the bearings or shaft of a pumpand temperature rise. This discus-sion applies specifically to singlestage, low specific speed (400-800rpm*√gpm/H3/4) mag-drive pumpsin the 1-30 hp range. Low flow oper-ations of higher specific speedpumps can lead to additional prob-lems such as suction recirculation,

which will not be discussed in thisarticle.

Radial Load on the Bearings and Shaft

When a standard centrifugalpump operates off its best efficiencypoint, the impeller experiences high-er radial loads due to hydraulicunbalance in the casing. The radialload becomes severe when thepump is operated near shut-off. In astandard sealed pump the loads onthe bearings are much higher thanthe load on the impeller. Typically,they are two times higher. This is aconsequence of the long overhangdistance between the impeller andthe first bearing, which is necessaryto provide adequate space for theseal.

In contrast, the first bearing of asealless pump is located very closeto the impeller. This results in bear-ing loads only slightly greater thanthe impeller loads. Figure 1 showstypical mag-drive pump bearingarrangements. The dimension Ldenotes the span between theimpeller and the first bearing. Eachlayout has its own strengths andweaknesses, but all provide for abearing close to the impeller. InType 3 the bearing rotates with theimpeller and consequently is veryclose to the load. With this designthe overhang distance, L, canapproach zero. The Type 4 designalso has the bearing close to theimpeller, but it uses a shaft can-tilevered from the containmentshell. This design is typically usedonly in small pumps because of thestresses at the shaft-to-containmentshell connection.

CENTRIFUGAL PUMPS

HANDBOOK

Photo 1. Example of an industrial non-metallic magnetic drive pump

PHOT

O CO

URTE

SY O

F AN

SIM

AG, I

NC.

Page 166: Centrifugal Pumps Handbook

The Pump Handbook Series 167

be manufactured from materialssuch as pure sintered silicon carbidethat effectively provide zero wearfor the life of the pump.

Temperature RiseThe temperature rise (ºF) of the

liquid passing through a pump isgiven by

∆T= H 778˙Cp˙η

where H is head in feet, Cp isspecific heat in Btu/lb ºF, and h isefficiency, written as a decimal val-ue. This equation assumes that alllosses result in heat that remains inthe liquid. To predict the tempera-ture rise in the fluid accurately, theefficiency factor in this equation

must include all losses. The threetypes of efficiency loss in magneticdrive pumps are: (1) hydraulic, leak-age and friction losses, (2) radial andthrust bearing friction, and (3) eddycurrents in the containment shell(rear casing).

(1) Hydraulic losses areinherent in all centrifugalpumps. However, the efficiency ofimpellers is improving with betterdesign, and, with the use of morepowerful magnets, the size of theinner magnet assemblies and theirresultant fluid friction losses areshrinking. Remember, though, thatsince the efficiency will always bezero at shut-off, the temperature risewill rapidly increase as the flow rateapproaches zero.

(2) Radial and thrust bearingfriction account for the smallestportion of efficiency losses. Forexample, a 1 1/2 x1x6 ANSI pumpwill lose 0.1-0.26 hp to bearing fric-tion. This represents only 3-9% ofthe shut-off power of the pump.

(3) Eddy current losses. Inmetallic magnetic drive pumps thecontainment shells are usually madeof a nickel alloy (e.g., Hastelloy®) orstainless steel. Both are electricalconductors. These stationary shellsare placed between the two sets ofrotating magnets within their pow-erful magnetic fields. When a mag-netic field moves past a conductorsuch as a containment shell, eddycurrents are generated. Generally,the eddy current loss for a 0.060”A mag-drive pump, therefore,

has a significant advantage in termsof impeller deflection, and it is moreresistant to the radial loads encoun-tered during low flow operation. Forexample, an impeller deflection of0.005" can be a problem for seal lifebut is not a concern in most seallesspump designs.

It is important that mag-drivepumps operating at low flows bedesigned with bearings able to han-dle consistently higher radial loads.The design must also provide foradequate cooling and lubricationflow to the bearings at these lowflow rates. One reason for theincreasing acceptance of mag-drivepumps in low flow rate service isthat product lubricated bearings can Photo 2. A Zirconia containment shell

Impeller and magnet a single unit. Stationary cantilever shaft. Bearings rotate with impeller.

1. Simple design2. Non-clog fluid paths.3. Containment shell (rear casing) supports shaft

bending moment. Shaft socket in containment shell must be very strong and resist creep if non-metallic.

• Black squares represent the mounting locations of the stationary bearings (type 1 and 2) or the mounting locationsof the stationary shaft (type 3 and 4).

• Red rectangles represent product lubricated bearing locations (type 1-4).

Impeller and magnet separate, connected via rotatingshaft. Bearings stationary and supported in a bushingsupport.

1. Front bearing has higher load capability than type 4.2. High manufacturing cost and more complex than

type 3 and 4.3. Popular in metallic pumps.

Impeller and magnet separate, connected via rotatingshaft. Stationary front bearing behind impeller.Stationary rear bearing in containment shell.

1. Front bearing has higher load capability than type 1.2. Longer span between bearings than type 1, 3 and 4.3. More chance of dry running or vapor lock at rear

bearing, since difficult to lubricate.

Impeller and magnet a single unit. Shaft is stationary andsupported at both ends. Bearings rotate with impeller.

1. Front bearing has highest load capability of all 4 types.

2. Simple design3. Lower stress concentration on plastic parts than

type 4.4. Non-clogging fluid paths.5. Popular in non-metallic pumps.

Basic Construction Description

L

L

L

L

Type 1

Type 2

Type 3

Type 4

Page 167: Centrifugal Pumps Handbook

168 The Pump Handbook Series

thick nickel alloy containment shellis about 15% of the magnetic cou-pling rating. If a 10 hp coupling isused, about 1.5 hp or 1.1 kW ofpower is directly transferred to theliquid at the containment shell. Thiswould be equivalent to equippingthe pump with a 1.1 kW heater. Additionally, theheat generated in the containmentshell remains essentially constantregardless of pump flow rate. This1.1kW can increase the temperatureof water flowing at 1 gpm byapproximately 7.5ºF. Cooling of thecontainment shell and prevention offlashing are the primary constrain-ing factors for metallic mag-drivepumps at low flow rates.

Performance curves for a non-metallic mag-drive pump are shownin Figure 2. The efficiency curveincludes all losses. As with all pumpperformance curves, the efficiency iszero at zero flow. If we take valuesfrom the efficiency and TDH curvesand insert them into the temperaturerise equation, given above, we cancalculate the temperature rise forthis pump. Figure 3 illustrates thetemperature rise for flow rates of 0-20 gpm. If a maximum temperaturerise of 10ºF is specified, then thispump can be operated at 2 gpm forwater. Slightly higher minimum flowrates may be required for other liq-uids due to their lower heat capaci-ties. Figure 4 provides the minimumcontinuous flow rates for a series ofthree non-metallic mag-drive pumps.These flows are based on a tempera-ture rise limit of 10ºF for pumpingwater.

Dry and Semi-dry Running Dry running and semi-dry run-

ning are the most common causes offailures in magnetic drive pumps.Damage can be caused by excessiveheating in metallic containmentshells and/or by mechanical or ther-mal shock at the bearings and shaft.

If a metallic containment shell isused and a pump runs dry, the shellwill be rapidly heated by eddy cur-rents, with temperatures rising to near-ly 1000ºF. This temperature rise is soquick that even jogging the pump (e.g.,to check rotation) is not recommended

K1516 Performance Curve3500 RPM, S.G. 1.00

0 50 100 150

1

2

3

4

5

6

7

8

9

10

0

20

40

60

80

100

120

140

160

180

0

200

Capacity (U.S. Gallons)

TDH

(Fee

t)

Effic

ienc

y (%

x 1

0)Po

wer

(hp)

Temperature Rise in a K1516 Pump @ 3500 rpm25

20

15

10

5

00 5 10 15 20

Flow Rate, gpm

Tem

pera

ture

Ris

e, °

F

Water

50% Sodium Hydroxide

37% Hydrochloric Acid

98% Sulfuric Acid

Figure 2. Performance curves for ANSI 1 1/2 x 1 x 6 size non-metallic magnetic drive pump

Figure 3. Temperature rise in same ANSI pump

0 2 4 60

100

200

300

400

500

108

Capacity (U.S. Gallons)

TDH

(Fee

t) ContinousOperation

Region

CP = 1.0 Btu/lb°F(Water)

Figure 4. Minimum continuous flow for non-metallic magnetic drive pumps

Page 168: Centrifugal Pumps Handbook

The Pump Handbook Series 169

when there is no liquid in the pump. Afew minutes would be long enough todemagnetize the magnets completelyand ruin the shell.

Non-metallic containment shellsgenerate no eddy currents and there-fore no heat, and so do not experiencethis kind of failure. Some manufactur-ers provide non-metallic containmentshells for their metallic pumps. (Photo2.) Typical materials of constructionare zirconia ceramics or plastics.

The critical component withrespect to dry running in pumps withnon-metallic containment shells arethe bearings. The bearings aredesigned to operate while wetted bythe pumpage and will exhibit a verylow coefficient of friction in this con-dition. Some material combinationssuch as a carbon bushing on sinteredsilicon carbide will maintain a rela-tively low coefficient of friction evenwhen dry. Such combinations arethus more forgiving in instances ofdry running. Sintered silicon carbideagainst sintered silicon carbide is thebest bearing material when wet, butit will typically show a suddenincrease in friction level when bonedry.

The resultant increase in shafttemperature is shown in Figure 5.The temperatures were measuredusing a non-metallic mag-drive in abroken suction application. Since thecarbon/graphite bushing has a lowercoefficient of friction when dry, thepump shaft temperature increasegradually. The silicon carbide bush-ing exhibits the same temperaturerise in the beginning, but then analmost instantaneous temperaturerise will occur. Both material combi-nations provide plenty of time to shutdown the pump before damage

occurs provided proper monitoringequipment has been installed.

During dry running the pump bear-ings and shaft will become hot. If whilehot, cool liquid is reintroduced to thepump, the bearings and shaft may faildue to thermal shock. Different ceramicmaterials have far different thermalshock limits. For example, aluminaceramic can survive only a 200ºF ther-mal shock while sintered silicon carbideis safe up to 600ºF. Silicon carbide canresist thermal shock because it has avery low coefficient of thermal expan-sion combined with a very high thermalconductivity. This allows the material toequalize in temperature very quicklywhile undergoing very small thermalstrains. Some manufacturers provideother bushing/shaft combinations. How-ever, in an actual service such as unload-ing, the coefficient of friction isunpredictable. In a laboratory test, thecarbon bushing/silicon carbide shaftcombination allows extended periods ofdry running if the bushing is new andclean, or if the pump suction and dis-charge are open and air flows freely tocool the shaft. In practical operation, dryrunning is not recommended in magnet-ic drive pumps because a) the majorityof magnetic drive pumps are still usingmetallic containment shells, b) residueon the shaft and bushing from a previ-ous pumping operation can change thefriction coefficient, c) abrasive matter inthe fluid can alter the bushing and shaftsurface finish, d) vibration from otherequipment may match natural frequen-cy of the pump rotating parts, and e) pip-ing usually restricts cooling by air flowthrough pump.

Monitoring, which is strongly rec-ommended, can include a pressureswitch, flow switch, electrical currentmonitor or electrical power monitor. A

pressure switch or flow switch is reli-able as long as liquid is clean and extraelectrical wiring in the pump field isfeasible. An electrical current monitoror electrical power monitor is very pop-ular since no extra wiring into thepump field is required and the devicecan be easily placed outside the haz-ardous area. Exercise caution if themotor is selected for the pump’s maxi-mum requirements but actual opera-tion is at a very low flow rate. When amotor runs at less than 50% of ratedload, current monitoring will not besufficiently sensitive. This is due to thecharacteristics of a motor. However,motor input power remains sensitivebelow 50% of rated load. For this rea-son sensing the performance changesbetween a pump operating at a lowflow condition and the same pumprunning dry will require a power mon-itor. These devices can be convenient-ly installed at the motor starter box.

However, power monitoringmay not be sufficiently sensitive todetect the changes between very lowflow operation and shut-off. Thesesituations will require the use of aflowmeter or differential fluid tem-perature measurement between thepump inlet and discharge.

CONCLUSIONIt is essential for pump users to

determine if the minimum flow ratespecified by the pump manufactureris constrained by a radial load on animpeller or by heat rise. In the caseof a radial load limitation, the specif-ic gravity of the pumped liquid mustbe taken into account. If heat rise isthe major limiting factor, the specificheat of liquid must be taken intoconsideration. These two factors areunrelated. ■

Kaz Ooka is President and co-founder of Ansimag, Inc. He has a B.S. inMechanical Engineering from Tokyo Den-ki University. Manfred Klein is ChiefDesign Engineer for Ansimag, Inc.. Heholds a Bachelors and Masters degree inMechanical Engineering from CarletonUniversity in Ottawa, Canada.

Time (min)0 5

50

100

150

200

250

10 15 20 25 30 35

Tem

pera

ture

(F°)

SiC Bushing Carbon Bushing

Figure 5. Shaft temperature U.S. time

Page 169: Centrifugal Pumps Handbook

170 The Pump Handbook Series

Sealless Options Optimize SolutionsIf zero leakage is the goal, sealless pump options can help tailor-make the solution.By Robert C. Waterbury, Senior Editor

Leak free? Zero emissions?Hermetic sealing? Whenenvironmental protec-tion needs and hazardoussubstances raise ques-tions, sealless centrifugalpumps often provide theanswers. The term "seal-

less" generally describes a class ofpumps that do not allow fluid leak-age into the environment. Andalthough this de-scription covers anumber of pump types, the two mostprominent examples are the cannedmotor pump (CMP) and the magneticdrive pump (MDP).

According to David Carr, mar-keting specialist at Sundstrand FluidHandling, neither the CMP nor MDPrequires a dynamic shaft seal to con-tain the pumped fluid. Instead, a sta-tionary containment shroud isolatesthe pumpage from the ambient envi-ronment. In the case of the magneticdrive pump, power is transmittedacross the stationary shroud throughmagnetic lines of flux. These inducerotation in the impeller shaft andthus avoid the leakage that is a nor-mal byproduct of mechanical faceseals.

MAGNETIC DRIVE PUMPSMagnetic drive pumps are sim-

ply centrifugal pumps with an inte-gral magnetic coupling between thedriver and the liquid end. The mag-netic coupling replaces the sealchamber or stuffing box so that theliquid end is hermetically sealed.The mechanical seal or packing iseliminated, and the only seal is a sta-tionary gasket or O-ring.

The two main subgroups ofmagnetic drive pumps are the syn-chronous and eddy current designs.

The synchronous type typically usesrare earth magnets. Because thesecan be adversely affected by tem-peratures in excess of 400ºF, specialauxiliary cooling provisions areoften required for such applications.

The eddy current designs, ac-cording to the Kontro Co. of Sund-strand Fluid Handling, employ atorque ring that is normally unaffect-ed by temperatures experienced inhot oil heat transfer systems. Eddycurrent sealless pumps feature arotating assembly that is sealed by acontainment shell. Power is transmit-ted through permanent magnetsmechanically coupled to the driver,rather than through motor windings.Cooling water is not requiredbecause the outer magnets andantifriction bearings are remotelylocated.

A torque ring integrally con-nected to the impeller shaft formsthe inner rotating element. Thisassembly is supported by journalbearings that are lubricated usingrecirculating hot oil. No precooling isrequired. The recirculated flow alsoremoves the heat generated from themagnetic coupling losses.

CANNED MOTOR PUMPSThe canned motor pump con-

sists of an induction motor whoserotor is integral with the impellershaft. A thin metallic can mechani-cally separates the rotor from thewindings and seals the pumpagefrom the stator. The rotor is support-ed by journal bearings that are lubri-cated by recirculated hot oil.Recirculation is provided by anexternal tube that feeds hot oil intothe back end of the pump.

CMP features may include a

two-bearing single shaft arrange-ment, dry stator and sealed junctionbox, both primary fluid containmentand secondary leak containmentshell, a controlled bearing operatingenvironment with monitor and aminimum of required components.

The benefits or advantages ofusing CMPs include: no shaft sealsand no external leak paths, no bufferpots, no buffer or process fluid leak-age disposal, no coupling or align-ment problems, low noise levels andlow maintenance costs.

Because high temperatures rou-tinely encountered in hot oil sys-tems normally exceed the limits ofthe motor, CMPs are designed withan integral cooling water heatexchanger. This exchanger sur-rounds the outside wall of the statorand removes the heat associatedwith motor losses. The recirculation

CENTRIFUGAL PUMPS

HANDBOOK

Photo 1. Chempump NC Series cannedmotor pumps feature an electronic diagnostic system.

Page 170: Centrifugal Pumps Handbook

The Pump Handbook Series 171

consist of welded primary contain-ment and hermetically sealed sec-ondary containment. For efficiencyreasons few MDP manufacturersoffer secondary containment.

Most MDP manufacturers,however, do offer secondary controlin the form of mechanical seals onthe OMR shaft penetration. Also, anMDP typically provides a thickercontainment shell. This offers moreresistance to corrosive or mechani-cal penetration. CMP primary linerthicknesses normally range from0.022"-0.035". MDP containmentshells range from 0.029"-0.060".

These differences are offsetsomewhat by the ability to monitorbearing and internal rotor positions.CMPs, by design, are easier to mon-itor than MDPs. This means thatthey are perhaps more apt to detectextreme bearing wear prior to con-tainment shell contact or penetration.

Solids/slurry handling. Stan-dard CMP and MDP products willaccommodate moderate amounts ofsolids. Optional designs for both arecapable of handling higher slurryconcentrations. With a slurry modi-fication, CMPs can deal with con-centrations in the range handled bymost standard centrifugal pumps.The MDP, however, cannot be iso-

lated as easily as a CMP and thusrequires a different internal bearing-mag coupling flow path than offerednormally.

Heat input. Heat is added byhydraulic and drive inefficiencies ineither type of sealless pump. For ourpurposes we can consider thehydraulic efficiencies to be the samefor both types.

Thus, for CMP design, a highefficiency motor is 80-85% efficient.But due to the ease in isolating themotor area, CMPs can offer optionalconfigurations to control the fluidtemperature, pressure or both toprevent product vaporization. How-ever, because a CMP motor tends tobe one large insulated mass once theunit is shut down, it allows heatsoak to occur. This permits theprocess fluid to be heated to highertemperatures at potentially lowerpressures than during normal opera-tion. The result could lead to flash-ing of the contained fluid. Also, thepump may vapor lock if restarted.

Synchronous MDPs are knownto be adversely affected by tempera-tures above 400ºF. But a synchro-nous drive is also more efficientthan an eddy current drive. A non-metallic shell is more efficient thana metallic shell. The synchronousdrive MDP with a metallic shell hasa drive efficiency of 80-85%, whichis comparable to the CMP. Howev-er, it lacks the insulated massaround the containment shell and isless susceptible to heat soak.

Cooling requirements. BothCMP and MDP can be configuredfor operation at elevated tempera-tures. Different designs require tem-perature limitations based uponspecific components and thereforemust be evaluated on a case-by-casebasis.

Jacketing. Steam or hot oiljackets can be added to either typeof pump to maintain the properproduct temperature. This ensuresthat the pumpage will be liquid at alltimes. CMP designs allow jacketingof the pump case, stator and rearbearing housing. Most MDPs canhave jacketed casings that add heatin the area of the containment shellwithout fully encapsulating it.

tube also makes a single pass insidethe shell of the exchanger. Thus, therecirculated pumpage is precooledto a safe temperature before enter-ing the back end of the pump onceagain.

CMP AND MDP FEATURES ANDOPTIONS

The first and most obvious jobin selecting a sealless pump is todetermine the basic system require-ments and operating parameters.Operating criteria such as high orlow temperatures, fluid capacitiesin gpm, system pressures, abrasiveslurries and corrosive agents willdetermine not only the size andcapacity of the sealless pump, butalso any special constructionrequirements.

Once these criteria are identi-fied, it is possible to look at the com-parative features and optionsoffered by canned motor pumps andmagnetic drive pumps. Much of thefollowing material is adapted frominformation compiled by Sund-strand Fluid Handling.

Containment. This is a majorconsideration if there are healthand/or safety concerns. SeveralCMP manufacturers offer some typeof double containment. This might

Figure 1. Today's generation of magnetic drive pumps offers advanced features andmaterials of construction, as shown in this example of Kontro's A-Range design.

Silicon CarbideBearing withCarbon GraphiteOption

Replacable CasingWear Rings

Single CasingGasket

Hastelloy “C”Containment Shell

Fully ContainedMagnets

Dual Power- EndPull Out Design

Bearing Isolator

Synchronous Drive DesignUtilizing Samarium CobaltMagnets

Optional Secondary Containment

Replaceable Non-SparkingBump Ring

Optional INSIGHT SIGHT ®®

Bearing

Monitor and Protection Systems

NOTE: Liquid end internals utilizeKontro’s unique cartridge systemallowing fast, easy maintenance andminimum downtime.

Page 171: Centrifugal Pumps Handbook

172 The Pump Handbook Series

High suction pressure.Optional designs are available forboth CMPs and MDPs in applica-tions where the suction pressureand maximum allowable workingpressure requirements exceed stan-dard pressure design capability.

Modifications in the CMP designfor high pressure applicationsinclude the use of primary contain-ment shell backing rings, thicker sec-ondary containment shells,additional pressure containment bolt-ing and high pressure terminalplates. The double containment fea-ture of CMPs is maintained even forhigh pressure.

Modifying the MDP for highpressure applications includes use ofa thicker containment shell andadditional pressure containmentbolting. Typically, the CMP is moreefficient than the MDP in high pres-sure applications due to its increasedprimary containmentthickness.With thicker primary con-tainment, the MDP shows greaterhysteresis losses (inefficiency) thanthe CMP.

ANSI conformance. Althoughsome CMPs are available with ANSIdimensions or hydraulics, most arenot. Many MDPs, however, offerANSI dimensions and hydraulics.Further, most manufacturers ofsealed and sealless ANSI pumps offerinterchangeability between theirpumps' wet ends and bearing frames.

Efficiency. CMP wire-to-waterefficiency is defined as hydraulicefficiency times motor efficiency. Asmentioned in the discussion of heatinput, a typical CMP motor is 80-85% efficient.

MDP wire-to-water efficiency isdefined as hydraulic efficiencytimes motor efficiency times cou-pling efficiency. Again, the magnet-ic coupling is 80-85% efficient.However, containment shell metal-lurgy (or lack thereof) and magneticcoupling type play major roles incoupling efficiency.

Wet end hardware, however,further complicates the efficiency dis-cussion. Impeller and casing geome-try play a vital role in hydraulicefficiency. A Barske design (openradial blade impeller and diffuser

discharge) is typically more efficientin low flow/high head hydraulics. AFrancis design impeller, enclosedwith backswept vanes and anincreasing radius volute, is moreefficient at moderate-to-high flowswith low-to-medium heads. Theuser, therefore, must thoroughlyevaluate wire-to-water efficiency foran accurate comparison.

Internal clearances. Clearancesbetween bearing ID and mating sur-faces are typically the same (0.003-0.007"), although differences betweenother rotating parts do occur. TypicalCMP clearances between the rotorand stator liners vary by manufactur-er from 0.018-0.044" radially. MDPclearances between the inner mag-netic ring and containment shell areusually 0.030-0.045" radially. Thislarger clearance gives the MDP theadvantage of allowing more bear-ing wear prior to containmentshell contact.

Bearing monitoring. Althoughdifferent monitoring methods areavailable for MDPs and CMPs, thedesign of the CMP lends itself morereadily to real bearing monitoring.A CMP bearing monitor can pro-vide axial, radial and liner corro-sive wear indication. Bearingmonitoring features vary by manu-facturer, and care must be exer-cised in selecting a sealless pumpmanufacturer with the desiredmonitoring features.

Bearing material options. Mostsealless manufacturers offer a hardbearing – typically silicon carbide –running on silicon carbide or tung-sten carbide. Alternatively, a softbearing is normally carbon graphiterunning on stainless steel. Becausethe materials are the same, themounting and lubrication plansassume greater importance.

Number of bearings. CMPshave two bearings; MDPs have six(including the motor). Most expertsfeel that fewer is better. However, aproperly designed, applied and oper-ated MDP will last just as long as aproperly designed, applied and oper-ated CMP.

Noise. CMPs have no motor fanand thus produce less noise.

Space. A CMP with its integralpump and motor occupies less real

estate (has a smaller footprint) than acomparable MDP with its pump,coupling and motor. However, closecoupled MDPs that may require nomore space than a CMP are avail-able.

Orientation. CMPs can bemounted vertically or even hung ona pipe. Few MDPs can be mountedvertically.

Temperature fluctuations.Both CMPs and MDPs can effec-tively handle wide temperaturevariations in the pumpage. Manufac-turers should be consulted in evalu-ating specific operating conditions,however.

Initial cost. For general dutyservice (clean, cool, non-volatile) anMDP is usually lower in price than aCMP. As the application becomesmore tortuous and difficult, howev-er, pricing reaches parity and mayeven favor the CMP.

Installation cost. A standardCMP or close coupled MDP with noauxiliaries (coolers, flush systems,etc.) will have lower installationcosts than a frame mounted MDP.This is due to the smaller footprintand minimal foundation require-ments.When optional CMP configu-rations are considered, the cost ofcooling lines, reverse circulationlines and flush lines may equal thecost of an MDP. Instrumentationalso adds cost to both types of instal-lation, so situations must be evaluat-ed individually.

Simplicity. Although the num-ber of pump parts may vary by man-ufacturer, CMPs generally havefewer parts. As optional configura-tions and more parts are added,however, they can easily equal thenumber of parts in an MDP. Itemsadding parts and complexity includeauxiliary impellers, tilting washersand heat exchangers.

While some MDP manufactur-ers may add such equipment, mostoffer only two basic varieties: dis-charge-to-suction recirculation anddischarge-to-discharge recirculation.Adding a standard electric inductionmotor as well helps some usersaccept the sealless technology.

Field serviceability. CMPshave bearing monitors to predict the

Page 172: Centrifugal Pumps Handbook

The Pump Handbook Series 173

need for bearing changeout prior tocontainment breach. CMP bearingsare typically easy to replace.

MDPs have thicker containmentshells to prevent breach. However,they have limited ability to monitorlubricated bearings to determinewhen routine maintenance isrequired. Ease of bearing replace-ment also varies with manufacturer.

If a primary containment shell isbreached, even the most serviceableCMP must be sent outside to be"recanned." CMP manufacturershave countered by offering sparerotor/stator sets at 60-80% of the costof a new pump.

A containment shell breach ofan MDP usually results in the pur-chase of a spare process-wettedrotating assembly and a containmentshell. Pricing is typically 60-80% ofthe cost of a new pump.

Coupling alignment and vent-ing. The typical CMP design has theimpeller mounted directly on themotor shaft inside the containmentarea. Coupling and coupling align-ment problems are therefore nonex-istent.

Most MDP installations useframe mounted motors that requirecoupling alignment. Many MDP sup-pliers, however, offer close-coupleddesigns that nearly eliminate cou-pling alignment problems.

Both MDP and CMP designs aretypically self-venting back to theprocess piping and do not requireadditional external lines.

PRODUCT OFFERINGS Because sealless pumps are

often used in severe duty applica-tions, monitoring and diagnosticoptions are important aspects ofmany installations.

Monitoring and diagnostics.A monitoring and diagnostic systemcalled IntelliSense from CraneChempump displays the position ofthe entire rotating assembly in itsNC Series of canned motor pumps.The system provides precise, real-time wear data with an accuracy of0.001" radial and 0.002" axial. Thisinformation enables users to plansimple parts replacement longbefore costly failure occurs. The

entire diagnostic system is isolatedfrom the process fluid and is there-fore not a sacrificial part thatrequires replacement. The displayunit can be hand held, mounted nearthe pump at eye level for easy view-ing, or located in a remote controlstation.

Teikoku canned motor pumpsoffer the Teikoku Rotary Guardian(TRG), which is available either as abuilt-in meter or as a remote panelmeter. It monitors the running clear-ances between the stator and rotor,bearing condition and rate of wear,reverse rotation, loss of phase andshort circuit conditions.

Sundyne offers a remotemechanical bearing monitor. Itallows remote alarm or shutdown ofthe pump signalling a need for bear-ing maintenance to protect themotor from damage. A dry opera-tion protection meter is also avail-able. In services such as tankunloading, the meter detects lowload in time to shut down the motorto prevent dry operation. A similarsystem called INsight from Kontroprovides radial and axial bearingwear monitoring in its line of mag-netic drive pumps. Kontro alsooffers a port for thermocouple orRTD temperature monitoring, anamperage monitoring system to pro-tect against cavitation or dry run-ning, a sensor port for vibrationmonitoring as well as liquid andpressure sensing options.

Jacketing. Complete or partialjacketing of the pump case, motorstator and rear bearing housing isoffered by Sundyne to control tem-perature when heating or cooling isrequired. Teikoku and other manu-facturers offer certain pump lineswith built-in heat exchangers as wellas motor cooling jackets for hightemperature applications.

Crane Chempump offers threelines of high temperature CMPs:

• GH – external circulation for fluid temperatures to 650ºF without liquid cooling

• CH – internal circulation forfluid temperatures to 650ºF withoutliquid cooling

• GT – motor isolation with ex-

ternal cooling for fluid temperaturesto 1000ºF

The motors in the GH and CHmodels incorporate a high-tempera-ture insulation system. The motor inthe GT model employs an integralliquid-cooled heat exchanger.

Inducers. Teikoku, Sundyne,Crane Chempump and other manu-facturers offer a wide selection ofinducers to meet net positive suctionhead (NPSH) requirements. It is notunusual, according to Teikoku, forone of its pumps to operate at a spe-cific suction speed of more than13,000.

Casings and impellers. Singleand double volute casings areoptions used in CMPs from BuffaloCan-O-Matic and other manufactur-ers, depending upon pump size andservice requirements. Likewise, openand closed impeller designs can bespecified as options on many modelsdepending upon operating require-ments.

Bearings. In addition to variousmanufacturer options, there are spe-cial bearing constructions to takeinto account. Buffalo Can-O-Matic,for instance, features spring-loaded,self-adjusting and self-lubricatingtapered carbon graphite motor bear-ings. These are designed to distrib-ute and automatically compensatefor bearing wear. This maintainsconcentric rotation of the rotor-impeller assembly, preventsmechanical contact between therotor can and stator can, andimproves both operation and main-tenance. Segmented bearings of self-lubricating carbon gra-phiteconstruction are used on large Can-O-Matic pump motors. Again, theyare spring loaded to compensate forwear.

Other construction options.Most manufacturers offer optionalconstruction materials for certaincomponents. Buffalo Pumps, forinstance, provides casings in achoice of ductile iron or 316 stainlesssteel. Likewise, impellers are avail-able in a choice of cast iron, bronzeor 316 stainless steel according to therequirements of the application.

Chempump lists hardened rotor

Page 173: Centrifugal Pumps Handbook

174 The Pump Handbook Series

journals, pressurized circulation sys-tems, sealless junction boxes andexplosion-proof CMP designs amongthe options in its G Series line. AndTeikoku provides a large numberof adapters to accommodate manydifferent pump and motor combi-nations.

DRIVING FORCESEnvironmental protection and

zero emissions of hazardous sub-stances continue to drive the use ofsealless CMP and MDP technology.

Add to that, however, the philoso-phy of continuous process improve-ment and reduced maintenance, andthere is ample reason to expect thatsealless pump technology has abright future. It is appropriate notonly for new installations, but alsofor many retrofit applications. Andstandardized ANSI dimensions canhelp make sealless technology ahighly cost-effective alternative aswell. ■

Page 174: Centrifugal Pumps Handbook

The Pump Handbook Series 175

Vertical Turbine Pumps Power PetrochemicalsVerticals are a popular choice for low NPSH applications, versatility of constructionand minimal floorspace requirements.By Herman A. J. Greutink

Petrochemical plantsneed water, whetherfrom local water sys-tems, deepwell pumps,rivers, lakes or oceans.The vertical diffuserpump normally plays amajor role in providing

service water, cooling water andoccasionally fire pump service. Thistype of pump is also used in oil fieldproduction as well as oil field pres-surization. Process fluids rangingfrom crude oil to liquified petroleumgas and other liquids (sulphur, forexample) are moved by verticalpumps in one or more stages ofextraction or production.

Frequently, liquified petroleumgas, propane, butane and anhydrousammonia are supplied from under-ground caverns.

In-plant pumps with very lowNPSHA are also apt to be the verti-cal turbine type. To economize onlength of barrel or can and reduceinstallation and pump costs, manyvertical pumps are supplied with afirst stage low NPSHR impeller.Some users require that the suctionspecific speed be limited to perhaps11,000. This limitation may be high-ly important for certain types ofpumps, but if a properly designedimpeller is used in a vertical turbinepump application, experience showsthat the suction specific speed canrange up to 15,000 without creatingproblems. Range of operation overthe pump curve and recirculationpossibilities must be considered.

SUPPLY WATER PUMPSFollowing are brief descriptions

of various vertical diffuser pumpsused in the petrochemical field:

Deepwell pumps (Figure 1)are commonly used to raise waterfrom underground aquifers. Line-shaft pumps are either oil lubricatedor water lubricated, and they arebuilt mainly to AWWA standards.Submersible motor driven pumpsare also used. Materials of construc-tion are mainly steel or cast iron forheads and bowls and bronze forimpellers.

Service water pumps (Figures2 and 3), used to pump from ponds,

lakes, rivers or oceans, are generallylarger than deep well pumps. Pumpsintended for fresh water intake havesteel columns and heads, cast ironbowls and bronze impellers. Forpumping brackish or sea water,coated standard materials are nor-mally used. Experience dictateswhether coated standard materialswill offer acceptable life. Otherwise,stainless steels (316, 316L, duplexstainless) or nickel aluminumbronze may be specified.

CENTRIFUGAL PUMPS

HANDBOOK

A 500 hp, 1800 rpm product pump operates in a midwest chemical plant.

PHOT

O CO

URTE

SY O

F JO

HNST

ON P

UMPS

, INC

.

Page 175: Centrifugal Pumps Handbook

176 The Pump Handbook Series

course, mechanical seals also needproper maintenance and installa-tion. But if either type of sealingtechnology is used and applied prop-erly, maintenance labor can begreatly reduced.

Vertical turbine pumps usedin re-pressurization of oil fieldspump produced water at high pres-sure – up to approximately 4500 psi.Because this water is sometimesvery corrosive, stainless steel bowlassemblies are commonly used, and

barrels and heads are made mainlyof coated steel (Figure 5).

There is an application in whichno water will show up at the surfaceand pressure is throttled off belowthe surface back to the water sup-ply. This method can be used whenpumping from atmospheric or vent-ed bodies of liquids.

PUMPING RAW STOCK AND PRODUCTMost pumps that transport

product or raw stock in the petro-chemical industry are built to API

Vertical cooling tower andplant supply water pumps (Fig-ure 4) follow fresh water materialstandards and construction. Line-shaft pumps for moving water arenormally built with a packing boxfor sealing. Lately, however, therehave been requests for water pumpsbuilt with mechanical seals. Perhapsthis is due to packing box mainte-nance requirements. There is an artto using packing boxes properly(they need to leak some water). Of

Driver

HeadAssy.

ColumnAssy.

BowlAssy.

Figure 1. Variation on water lubricateddeep well turbine pump

Driver

Motor Standand Tube

Tension NutAssy.

ColumnAssembly

With BelowGround

Discharge

BowlAssy.

Figure 2. Supply and drainage pump, axialflow (propeller) from 5'-20' of head

Driver

DischargeHeadAssy.

ColumnAssy.

BowlAssy.

Figure 3. Mixed flow type service water,plant water, with heads of 20-60' perstage

WaterLevel

Page 176: Centrifugal Pumps Handbook

The Pump Handbook Series 177

610. Many users add customrequirements.

In general, the specifications aretight, and extra attention to detailmust be paid to the pump's properdesign and fabrication. Most verticalturbine pumps in these applicationsare built as barrel or can pumps (Fig-ure 6) with mechanical seals. Pres-sure containment constructionconforms to ASME section VIII.

Sealing and bearing clearancespecifications also must be fol-lowed closely. The mechanical seal

configuration depends on the prod-uct pumped and in many casesmust adhere to strict environmen-tal protection rules. Most of thesepumps have motor drives withthrust bearings in the motor,although some requirements callfor the European style thrust bear-ing in the pump. The reason forthe latter is better mechanical sealmaintenance with less run-out.However, this increases total headand motor assembly, which couldaggravate vibration problems. Italso increases maintenance diffi-culty as more parts must be disas-sembled to get to the mechanicalseal.

Pipeline pumps may be horizon-tal multistage pumps, but boosters

from the tank farms to the pipelinepump are preferably vertical can orbarrel pumps (Figure 7). These offerbest utilization of storage capacity.The available NPSH can be very low– putting the first stage at a levelwhere the tankage can be pumpedempty to zero NPSH available and atthe same time supply the necessaryNPSH through the vertical barrel orcan type pump to the horizontalpipeline pump. High pressurepipeline pumps also can be builteconomically using vertical multi-stage pumps up to approximately2000 hp.

Extremely high or low tempera-ture liquid applications are designedto individual requirements. Forinstance, sulphur is handled by ver-tical turbine pumps with steam orhot oil jacketing from top to bottomof the pump. Hot oil circulatingpumps (for heat transfer) are built towithstand forces caused by temper-ature differentials and pipe loading.

The pump specifier shouldremember that material require-ments per API 610 are aimed mainlyat horizontal centrifugal pumps. Therequirement for cast steel is all rightfor most horizontal pumps, but forturbine pump bowls and impellers itcan create problems that are expen-sive to correct. A better way gener-ally is to make the bowls andimpellers out of stainless steel, espe-cially the 300 series and duplexstainless.

Practically all vertical turbinepumps are made to order. Off-the-shelf items consist only of standardbowls, impellers, cast iron headsand some smaller pieces like thread-ed shaft couplings. As a result, com-munications among user, consultantand manufacturer must be verygood to get the best equipment forthe application.

CONSTRUCTION OPTIONSAs mentioned, construction

materials vary according to fluidcharacteristics. Materials are oftenselected according to the corrosive-ness and abrasiveness of thepumped liquid. The following aresome construction variations.

Driver

DischargeHead Assy.

Barrel

BowlAssy.

Figure 5. Oil field pressurization pump.High pressure multiple pumps (barrel orcan) can be used in series.

Suction Discharge

PackingBox

Driver

DischargeHead Assy.

ColumnAssy.

BowlAssy.

Figure 4. Service water/cooling towerpump with heads from 50' and up perstage

WaterLevel

Page 177: Centrifugal Pumps Handbook

178 The Pump Handbook Series

Below-base discharge and/orsuction is one option to the moreprevalent above base-dischargedesign. The figures show variationsin packing boxes and mechanicalseals.

Product lubricated lineshaftbearings can be used as an option tooil lubricated or clean water flushedlineshaft bearings. Rubber bearingsand bronze bearings are used mostoften. Rubber is used successfully onhard shafting or hard-faced shaftingin water with abrasives such assand. Do not use rubber bearings inpetrochemicals that attack rubber.We also see the use of metal-filledgraphite bearings in light ends and

condensate pumps. Where abrasivesare present and rubber cannot beused, hard bearings (from nitronic60 to carbide) can be used on proper-ly selected shafting such as 17-4 pH,nitronic 50 or hard-faced shafting.

Cast iron bearings can be usedin oils. It is also possible to coat castiron bearings for corrosion protec-tion in water pumping applications.

Constructions shown in the fig-ures can be altered to suit therequirements of the system in whichpumps will be used. Turbine pumps

have been built to run upside down,horizontally or at an angle. The vari-ations of driver systems and con-struction systems are endless. Anytype of driver with right angle gears,vertical motors (hollow or solidshaft) and submersibles can be used.

Obviously, multi-staging allowsone to reach the head required at agiven capacity (it's like connectingpumps in series). However, com-plete pumps can be put in series, too.For instance, driver, constructionand shaft limitations of a 12" multi-stage pump may limit a pump to per-haps 1500 psi. If 4500 psi is needed,just put three pumps in series.

INSTALLATIONInstalling vertical turbine

pumps is a matter of keeping all thepieces or sections in good shape.First, all the pieces that have to fittogether to make a vertical turbinepump run properly must be manu-factured with three basics in mind:straightness (of shafting), concentric-ity and parallelism. This is true forall the pieces that fit together tobuild a total pump. If these consider-ations have been well addressed,one normally should be able to turnthe shafting by hand after theimpellers have been lifted off theseat. Once this condition has beenreached, the pump must be set toturn in the right direction when thepower is applied. Most pumps of thevertical turbine, mixed flow and axi-al flow type turn counter-clockwisewhen viewed from the top. But somevery old units may turn clockwise –so pay attention.

For pumps with threaded line-shaft, the motor/driver rotation mustbe checked before the driver ishooked up to the pump. In the caseof slipfits between motor and shaft,remove the piece that forms the slip-fit before checking rotation. Other-wise, the pieces may gall and causeserious and expensive damage.

Check the pump's performance,vibration and runout when it is putin service. At times, runout can alsobe checked before running. In mostcases the pump will operate satisfac-torily, but sometimes problems thatrequire immediate attention arise.

Driver

DischargeHead Assy.

Barrel

BowlAssy.

Figure 6. High pressure LPG pump (barrelor can)

Suction Discharge

MechanicalSeal

Driver

BowlAssy.

Figure 7. Pipeline or pipeline boosterpump

FlangedColumn

FlangedCoupling

MechanicalSeal

SuctionBarrelwithBelowBaseSuction

Page 178: Centrifugal Pumps Handbook

The Pump Handbook Series 179

Excess runout often can be cor-rected by changing coupling parts 90or 180º (where bolted couplings areused). It is more difficult to correct ifthe shaft is bent or some registers arenot true or some faces are not flat.Check driver base and shaft and gofrom top to bottom. Also, dirtbetween butting faces frequentlycauses major problems!

Vibration. Considerable runoutmay be accompanied by vibration. Ifthere is no runout problem, howev-er, it is necessary to determine whatcauses the vibration. First discon-nect the driver from the pump andrun the driver. If it does not vibrate,then check the pump. If the driverstill vibrates, do the following: pushthe stop button or (in case of geardrive) slow down the engine driveand check vibration while slowingdown. If vibration diminishes inline with speed, it is a balancingproblem with the driver. Get it re-balanced. If vibration immediatelydisappears when slowing down, it isprobably a critical speed problem,which may have several causes: 1) the design is such that the dri-ver/head assembly has a criticalequal or close to the operatingspeed, 2) pipe loading, 3) founda-tion loading. In the case of electricmotors, an electric unbalancebetween leads can cause vibration,but this is rare. Try to find out what2) and 3) are doing by just looseningthe anchor bolts a couple of turns.Many pumps exhibit 2) and 3) as

problems. If so, it may be necessaryto rework the piping connection. Inany case, if the vibration problem isnot easily solved, get the pumpmanufacturer involved.

Sometimes a piece of wood orother matter gets into the pump andcauses vibration. Maybe it can bebackflushed out. And no matterwhat, the sump and the piping mustbe clean from the start.

When vibration begins to showup after years of running, it's time topull the pump and replace bearingsand repair it as new. Keep in mindthe three musts: concentricity, paral-lelism and straightness.

In a particular installation offour identical barrel pumps, one unitpersistently vibrated. The motor andeven the pump were replaced withstill no improvement. The investiga-tion, however, identified that thegrout under the barrel flange wastotally inadequate for the vibratingpump. In other words, weaker sup-port put the pump in a natural fre-quency mode equal to the runningspeed. This has been seen in otherinstallations where pumps weremounted on beams or slabs that hap-pen to lower natural frequencies tothe pump speed.

When using variable speeds, thepump may be in a natural frequencymode at one of the speeds. If so, it isbest to block out that area of speedin the controls.

Operation. When pumps areselected, built and installed proper-

ly, they will have a reasonable lifeexpectancy. If a pump's life is not upto reasonable expectations, its selec-tion, use, construction and installa-tion must be thoroughly checked.Performance can be changed to suithead capacity requirements; con-struction and materials can beupgraded as necessary; and bearingsystems, including the thrust bearingsystems in the driver, can beimproved. Many vertical pumps willstart in upthrust. So, the bearing sys-tem must be built to handle that, andin some cases the pump may berequired to run in continuousupthrust. The system can be built forthat, too.

Sometimes particular pumpscontinually vibrate or rattle or runout and repairs have to be made fre-quently. That's when time must betaken to solve the problem properly,whatever the problem is – from flowto the pump, head, capacity and allmaterials and construction inbetween!! ■

Herman Greutink is Vice Presi-dent and Technical Director for John-ston Pump Company (Brookshire, TX).Among other professional affiliations,he is a longstanding member of TexasA&M's International Pump User’sSymposium Advisory Committee andthe Hydraulic Institute, and he fre-quently contributes to Pumps and Sys-tems on the subject of vertical turbinepumps.

Page 179: Centrifugal Pumps Handbook

180 The Pump Handbook Series

Chopper Pumps Digest The SolidsThese tough workhorses eliminate plugging problems in heavy-duty applications.By John Hayes

If the hydraulic performancecharacteristics are right, chop-per pumps can be used in anyindustrial or municipal appli-cation that involve pumpingsolids-laden slurries. They area cost-effective means of elimi-nating pump plugging prob-

lems and optimizing systemperform- ance. Before discussingspecific applications for chopperpumps, however, let's look first atthe design details that make chop-per pumps unique.

DESIGN DETAILSA chopper pump is a cen-

trifugal pump that uses sharp-ened semi-open centrifugalimpeller blades to cut against astationary bar across the fulldiameter of the inlet. This bar isknown as the "cutter bar," andthis style of chopping and pump-ing is known as "positive chop-ping." All incoming solids toolarge to pass through the impellerare chopped prior to entry intothe pump, thus eliminating anypossibilities of pump clogging.Other items critical to the successof a chopper pump are a severeduty seal-and-bearing system,hardened wear parts and a histo-ry of successful pump installationsby the manufacturer. Solids cuttingby the impeller and cutter baroccurs when:

1. The suction created by therotating impeller pulls material intothe center areas of the pumpimpeller. To reach the impeller cen-ter areas, liquid and the materialentrained in the flowstream mustpass through the cutter bar openingsin the lower suction plate locatedjust below (or ahead of) the impeller.

2. As material passes throughthe cutter bar openings into the lowpressure areas of the impeller, it gets

caught between one of the rotatingimpeller blades and one of the twostationary shear surfaces that arecast into the cutter bar. These sur-faces extend all the way across theintake opening in the cutter bar anddivide it into two segments. Positivechopping is required of all solids

entering the pump prior to pumping.3. The leading edge of each

impeller blade is sharpened andmachined flat on the blade facewhere it runs next to the cutter bar.This forms a cutting edge. The mate-rial caught between the impellerblade and cutter bar is severed.

The advantage of the chopper isthat pumping and chopping are inte-grated into one efficient system. Thechopping is done right where thepumping is done. The material natu-rally fits through the flow passagesof the impeller and casing. If the

material won't fit through, thepump keeps chopping it until itdoes. This is an example of positivechopping.

BENEFITSFollowing are some of the bene-

fits of positive chopping.

1. Positive chopping allowslarge, troublesome materials topass through the pump, elimi-nating downstream plugging ofvalves, heat exchangers, nozzlesor other pumps.

2. A chopper pump oftencan replace two pieces of equip-ment, a comminuter (or pre-grinder) and a "non-clog" pump.This approach is extremely cost-effective because the mainte-nance costs on comminutersalone can be very high.

3. Chopping material in thepump produces a more homoge-neous slurry and reducespipeline friction.

4. Chopper pumps can han-dle materials in sumps that noother pump can handle.

5. A severe duty seal-and-bearing system that incorporatesdouble row thrust bearings and amechanical seal reduces down

time by handling the heavy work-load of chopping and pumping solidsreliably.

6. Hard, wear-resistant pumpparts hold up to the rigors of chop-per pump service. Standard pumpimpellers and cutter bars made ofcast alloy steel heat treated to 550Brinell provide extended service lifein most applications.

7. Selection of a manufacturerwith extensive experience in severeduty applications helps assure theuser of dependable service. Chop-per pumps are not a commodity

CENTRIFUGAL PUMPS

HANDBOOK

Page 180: Centrifugal Pumps Handbook

The Pump Handbook Series 181

clog pumps. Chopper pumps elimi-nate plugging problems by choppingall incoming solids prior to pump-ing. Septage receiving pit pumping isa very tough challenge.

Because conventional non-clogpumps often require oversizing tohandle solids, another means ofpump protection is to pre-grind allmaterials with a comminutorupstream of the pump. This is anunnecessary added cost in installa-tion, operation and maintenancewhen a chopper pump, which is onepiece of equipment, does both jobs.

Sludge Transfer and Recircu-lation. One particular problem withsludge pumping is the "roping" ofhair and other stringy materials cre-ated by prerotation within the pip-ing ahead of non-clog pumps(especially vortex pumps). Chopperpumps eliminate roping by chop-ping the solids as they enter thepump. Another common problem ispumping grease-and-hair balls orother reformed solids. In digesterrecirculation, passing a grease-and-hair ball from the bottom to the topaccomplishes nothing. If thereformed solid is chopped duringthe recirculation process, then thechopped solids have a high surfaceto volume ratio and are digestedfaster.

Chopper pumps can handle ahigher solids content than conven-tional non-clog pumps. In normaltreatment plant processes thismeans that sludges can be more con-centrated without exceeding thepump's capability.

Chopper pumps also reducedownstream pipeline plugging andfriction losses associated with highsolids content. The chopping ofthese solids creates smaller solidswith sharp edges that tend to scourthe inner walls of discharge piping.Friction is reduced due to maintain-ing full pipe diameter, rather thanchoking flow with grease build-upon the pipe walls. Chopping andshearing of sludge tend to reduceviscosity, further reducing frictionloss.

Digester Scum Blankets. Bothaerobic and anaerobic digesters tendto form scum blankets when con-

ventional mixing methods are used.These blankets inhibit methane pro-duction and eventually require morefrequent cleaning of the digester. Avertical chopper pump with a recir-culation nozzle and sealed deckplate can be installed through anexisting manhole opening in the topof the digester and used to chop andmix the scum blanket. The object isto inject supernatant through anadjustable nozzle into the scumblanket approximately 1 foot belowthe surface. With the pump locatednear the periphery and the nozzleaimed at a proper angle to the wall,the action of the nozzle forces theblanket to rotate. Once the blanketis mixed, methane gas productionwill increase. The pump then can beused intermittently to keep the scumblanket mixed. This can increasedigester capacity up to 40% andincrease gas production up to 300%.

The action of chopping andconditioning material to reduceparticle size is of proven value insewage treatment plants. If diges-tible material is in smaller parti-cles, then the surface area of theseparticles is relatively large in com-parison to their volume. Bacterialaction can then be more effective

item. Every application is unique inits severe nature and duty. Procure-ment of inexpensive equipment orequipment that is "new" to the mar-ket could provide disappointingresults. Always ask the supplierabout the history of the productand its applications.

WASTEWATER APPLICATIONSLift Stations. Small residential

lift stations are equipped with sub-mersible grinder pumps, and largerlift stations with non-clog pumps.However, lift stations can experi-ence an unusually high concentra-tion of solids such as hair, rags orplastics that cannot be reliably han-dled by these conventional pumps.If heavy solids loading is anticipatedduring the engineering stage, manyof the larger lift stations include acomminuter ahead of each pump.Chopper pumps have solved manylift station plugging problems andeliminated the need for comminu-tors. In some instances chopperpumps have directly replaced exist-ing pumps without the need forrepiping.

Chopper pumps for lift stationscan be sized for hydraulic require-ments without regard to minimumsphere passing diameters. General-ly, non-clog pumps must be sizedaccording to maximum anticipatedsphere size, with hydraulic consider-ations secondary. For example, arequirement of 200 gpm at 130' tdhmight normally dictate a 3" pumphydraulically. But a 3" non-clogpump cannot provide the high headand meet a 3" sphere requirement atthe same time. Therefore, the engi-neer and/or manufacturer would beforced to use a 4" pump. This resultsin higher costs – due not only to thelarger pump, but also to lower effi-ciency and higher power consump-tion. However, because a chopperpump reduces the size of solidsbefore they enter the pump, sizingof the pump is based mainly onhydraulic requirements, with littleconsideration given to sphere size.

Septage Receiving. Septage iscomprised of concentrated solidsfrom septic tanks plus rags and plas-tics that can plug conventional non-

Photo 1. By slurrying the scum blanketwith a recirculating chopper pump, TonyKucikas of Nut Island WWTP rejuvenatedthe digester and saved Boston $1.5 mil-lion in cleanout costs.

Page 181: Centrifugal Pumps Handbook

182 The Pump Handbook Series

and rapid. Plants that have usedchopper pumps for digester recir-culation and/or scum blanket mix-ing have seen that particle sizereduction increases both the rate ofdecomposition of digestible materi-al and digester gas production.

Clarifier Scum. Inherentproblems with clarifier scuminclude plugging and air binding.The chopper pump addressesplugging problems. Air binding,on the other hand, must beaddressed indirectly.

Inherent recirculation aroundthe chopper pump inlet usuallycauses enough mixing to keep airbinding from occurring in scum pitswith short retention times. If ascum pit has a long retention time,then the scum may concentrate andform a blanket on the top. As thepit level is pumped down, this blan-ket can block the pump suction. Inthis case, a recirculation nozzleshould be used to pre-mix the scumpit prior to pump-out. Also,because chopper pumps haveheavy-duty oil bath lubricated bear-ing and seal systems, the scum pitscan be pumped completely down tothe pump inlet without damagingthe pump bearings or seal due toloss of coolant or lubricant. Thisallows full scum removal duringeach pump-down. By adding amotor low current monitoring sys-tem in the pump controls, the"OFF" function can be based onlow motor current draw rather thanan on "OFF" mercury float switch.Once the liquid level drops to thepoint where air enters the pumpinlet, the current draw will dropoff, and the low current relay willshut the pump off.

INDUSTRIAL APPLICATIONSFood Processing. Chopper

pumps can be applied in almost allfood processing waste handlingoperations. Applications includevegetable waste with whole vegeta-bles and stalks, poultry and turkeyparts with feathers and whole birds,beef processing with hair and flesh-ings, fish carcasses with offal, andany other food processing waste-water. Slurries are generally

pumped directly into trucks forhauling to feedlots or land applica-tion, sometimes first across a separa-tion system to reduce water content.

Wood Products. There aremany applications within the woodproducts industry where chopperpumps greatly reduce downtimedue to pump clogging. First, rawlogs are stacked in sorting yardswhere rainwater runoff sumps cancollect bark, limbs, rocks and largequantities of dirt. This requires aheavy duty pumping system, and isideal for a chopper pump. Next, thelogs might be debarked, leaving aslurry heavily laden with bark,most of which floats. Recirculationnozzle systems added to the chop-per pump help to suspend the barkwhile pumping. Then, the logs areeither sawn in a sawmill, orchipped in a pulp mill. Runoffsumps collect the chips and saw-dust, requiring further handling ofsolids too difficult for a standardnon-clog pump. Finally, waste pulpor broke is also easily handled by achopper pump.

Hydropulper rejects in woodproducts recycling are the most

severe pumping applications in allindustry. Rejected materials from therecycled bundles can contain wireand plastic strapping, large quantitiesof plastic sheeting or wrap and woodfrom pallets – all suspended in aheavy pulp slurry. Due to the heavynature of these slurries, chopperpumps are generally oversize, andmotor horsepower is increased tohandle the heavy chopping load.

Other Industries. The worldof heavy industry contains numer-ous other special applications forchopper pumps. The steel, chemi-cal, automotive, contractor, min-ing, sand & gravel, andpetrochemical industries all usechopper pumps in applicationsinvolving pumping of waste solidswith the occasional unknowns.Anytime a waste sump must dealwith the unplanned worker'sglove, pieces of wood crating,rocks, bottles, glass, cans, plasticsor other items larger than the sumppump is designed for, then a chop-per pump is applicable.

FINAL CONSIDERATIONSAll of the applications discussed

here center on the pump's ability tohandle solids from a pumping stand-point. However, we must not forgetthat chopper pumps can also elimi-nate seal and bearing failuresobserved in other pumps. Becausethe chopper design requires heavyshafting, an added benefit is longerlife resulting from stronger parts andless vibration. Quite often seal fail-ures in conventional pumps areassociated with solids wrapping orbinding at the impeller or seal. Thiscan cause severe vibration that istransmitted through the shaft andseals to the lower bearing. Thisresults not only in seal and bearingfailure, but also can introducemoisture into a submersible motor.The heavier shafting and shortoverhang of the chopper pumpbearing and seal design addressesthis problem and reduces mainte-nance costs as a result.

More industries, municipalitiesand engineering firms alike are dis-covering the economics of applying

Photo 2. A vertical chopper pump handlesbark, wood, rocks and dirt in a pulp milllog yard stormwater runoff sump.

Page 182: Centrifugal Pumps Handbook

The Pump Handbook Series 183

chopper pumps in applications inwhich conventional pumps have his-torically failed. These failures aregenerally due to plugging or seal-and-bearing failure, and all containhidden costs that must be addressed.The largest fallacy of the "low bid"system is that low initial price doesnot mean lowest overall cost. Moreoften than not, "low bid" equipmenthas a higher failure potential thanproperly specified and purchasedequipment. The solution starts withthe user's request to obtain equip-ment that will operate maintenancefree, and it ends with the foresight ofthose with purchasing authority tothink toward the future. ■

John Hayes has a BS inMechanical Engineering Technologyfrom Clemson University, has workedwith chopper pumps for over 12 yearsand is presently the MarketingManager for Vaughan Co., Inc.

Page 183: Centrifugal Pumps Handbook

184 The Pump Handbook Series

So You Need Pumps For A Revamp!Here are tips for specifying and selecting the right centrifugal pumps. By J. T. ("Terry") McGuire

Successfully specifyingand selecting pumps fora unit revamp requiresmany of the same disci-plines as for a new unit,but with a difference.The difference is that afull-scale working mod-

el is available for examination andanalysis. Taking advantage of thisopportunity can lead to employingpumps that consume less energy andhave a longer mean time betweenrepair (MTBR). This, in turn, lowersplant operating costs and can raiseplant output, hence revenue, throughhigher plant availability.

In a process unit, pumps moveliquid and raise its pressure to allowthe process to run, but their role isfundamental and their interactionwith the system a critical factor intheir performance. This last point,interaction with the system, leads tothe first step in specifying andselecting pumps for a revamp.

PUMP SYSTEM INTERACTIONA centrifugal pump operates at

the capacity determined by theintersection of its head capacitycurve and the system's head capaci-ty curve (Figure 1). At this point theenergy added by the pump equalsthe energy required by the system.Note in Figure 1 that the energyrequired by the system is oftenincreased by throttling across a con-trol valve to regulate the pump'scapacity. Also note that this meansof flow control is feasible only withkinetic pumps – those that add ener-gy by raising the liquid's velocity.

Displacement pumps (Figure 2),deliver essentially a fixed capacity ata given speed and thus add only theenergy needed to move that capacitythrough the system. Care is thereforeneeded to ensure that this energynever exceeds the mechanical capa-

bility of the pump. In other words,displacement pumps must always beinstalled with a full capacity reliefvalve upstream of the first valve inthe discharge system. And the reliefvalve must have an accumulationpressure (rise above cracking pres-sure to achieve full flow) that keepsthe pump's discharge pressure andcorresponding power below the max-imum allowable.

LEARNING FROM THE EXISTINGINSTALLATION

The system energy require-ments for new units are estimatedusing various assumptions and mar-gins. When centrifugal pumps areused, the system designer usuallyrelies on a control valve to balancesystem and pump energy at thedesired flow rate. In engineering arevamp, it is possible to determinethe actual system characteristicaccurately and thus avoid energy

losses due to conservative assump-tions and margins. In most casesthese losses are far greater thanthose resulting from differences inefficiency among various selectionsfor the same duty.

Determining the actual systemhead requires accurate measure-ment of:

• pump flow rate at one condition• pressure at the pump's suction

CENTRIFUGAL PUMPS

HANDBOOK

DISPLACEMENT

@ CONSTANTSPEED

KINETIC

FLOW

ENERGY ADDED

➞ ➞

Figure 2. Flow regulation of kinetic vs.displacement pumps

SYSTEM HEADw/ THROTTLING

SYSTEM HEADw/out THROTTLING

POWER ABSORBEDBY THROTTLING

NORMAL FLOW

POWER

HEAD

POWER

D

Z

2

3

V

V S

S

Figure 1. Centrifugal pump vs. system with control valve

Page 184: Centrifugal Pumps Handbook

The Pump Handbook Series 185

the points of suction and dischargepressure measurement. The frictionloss is significant when there areelbows, valves or reducers betweenthe gauge and the pump. The differ-ence in velocity head usually mat-ters only when the pump head islow and there is a difference of morethan one pipe size at the points ofpressure measurement.

Subtracting the static head com-ponents from the pump head, Figure4, yields the system friction head,HL1-4. When a control valve is usedin the system, the head being lost tothrottling across the control valve iscalculated from the measured valvepressure drop, then subtracted fromthe total system friction to give thehead lost to friction in the piping,including entrance and exit losses.

Recognizing that the head lost tofriction varies as the square of theflow rate, the equivalent system fric-tion at several other capacities nowcan be calculated and the systemhead characteristic plotted (Figure 4).If the static head varies with time, asit often does in a transfer process,then the range of system heads canbe plotted after allowing for maxi-mum and minimum liquid levels inthe suction and discharge vessels.

The other critical aspect of thesystem to be verified using themeasurements already made is theNPSH available at the pump. Formeasurements at the pump suction,

again referring to Figure 3, theequation is:NPSHA = (P2+ Pa- Pvap) 2.31 + Hz + Vs2(3) _________ ___

SG 2g

where Pa is atmospheric pressure atsite, Pvap is the vapor pressure of thepumped liquid at the pumping tem-perature, Hz is any correction forgauge elevation to the pump's refer-ence level, and Vs is the fluid veloc-ity at the point of suction pressuremeasurement. The pump's refer-ence level is the shaft centerline forhorizontal machines and the center-line of the suction nozzle for verticalmachines.

Because NPSH is also equal to:

NPSHA = (P1+Pa) 2.31 + Hz1 - HL1-2(4) _________

SG

it is possible with the measurementsalready made to calculate the frictionloss in the suction side of the system.And then following the same proce-dure used for the system head, thesystem NPSHA characteristic (Fig-ure 4) can be developed. If the liq-uid level in the suction vessel canvary with time, the range of NPSHAcan be plotted in the same manner asthe range of system heads.

With the net system head nowknown accurately, the head neededto move the required flow and allowflow control can be minimized. Atthe same time, an accurate NPSHAcharacteristic eliminates hidden mar-gins. This means an appropriateNPSHA margin can be set for theapplication, facilitating selection ofan optimum hydraulic design. Keep-ing the pump head to the minimumnecessary lowers energy consump-tion. And an optimum hydraulicselection can contribute to bothlower energy consumption andlonger MTBR.

Before preparing a pump speci-fication, two more factors must beaddressed at the site. The first is suc-tion piping. Many problems arecaused by poor suction piping. Aunit revamp is an opportunity tocorrect this. The important featuresof the suction piping layout are the

and discharge and in the suctionand discharge vessels

• liquid levels relative to a com-mon reference in both suctionand discharge vessels

• pressure drop across the controlvalve, if used, taking care tomeasure the downstream pres-sure some 10 diameters fromthe valve to avoid the influenceof any flow distortion

To make use of the pressuremeasurements, it is also necessaryto determine the pumped liquid'sspecific gravity (SG) at each measur-ing point. This can be determinedfrom liquid temperature as long asthe liquid being pumped is knownwith certainty.

Using Figure 3 as a reference,the system head is:

Hsystem= (P4 - P1)2.31 + (Hz2 - Hz1) + HL1-4(1) _________

SG

and the pump's total head is:

Hpump= (P3 - P2)2.31 + Hz + HL2-3 + ∆V2

(2) ________ ___SG 2g

where Hz is the correction for gaugeelevation, if any, HL2-3 the frictionloss between the suction and dis-charge pressure gauges, and ∆V2/2gthe difference in velocity heads at

➞ 1

4

3

2

P1

P2

DATUMLEVEL

Hz1

Hz2

Figure 3. Hydraulic gradient

Page 185: Centrifugal Pumps Handbook

186 The Pump Handbook Series

orientation of reducers, the proximi-ty of elbows to each other when indifferent planes, the orientation ofthe elbow immediately upstream ofdouble suction pumps, suction pip-ing slope and submergence over thevessel outlet.

When the suction piping rises to

the pump, reducers in horizontalruns must be eccentric and installedflat side up (Figure 6a). In suctionpiping from above the pump, hori-zontal reducers can be concentric oreccentric, installed either flat sideup or flat side down (Figure 6b). A

concentric reducer is necessary forend suction pumps of high Ns or S orboth. Eccentric reducers installed flatside down are used by many design-ers to eliminate low points in the pip-ing, which can accumulate dirt.

At normal suction piping veloci-ties of 7-8 fps, two elbows in serieswith their planes 90º apart shouldbe separated by 10 diameters andshould have the reducer downstreamof the second elbow (Figure 7). Sepa-rating the elbows in this mannerlargely dissipates the flow distortionproduced by the first elbow before itreaches the second. This avoidsdevelopment of a swirl at the outletof the second elbow. A reducerplaced downstream of the secondelbow helps dissipate the flow dis-tortion and any swirl that may havedeveloped.

The last elbow in the piping to adouble suction pump must be in aplane normal (at right angles) to theaxis of the pump's shaft (Figure 8a).An elbow in a plane parallel to thepump's shaft axis (Figure 8b) willcause uneven flow into the impeller.This can result in higher power con-sumption, noise, vibration, prema-ture erosion of one side of theimpeller and thrust bearing failure.

Moving away from the pump,the suction line must not have anyhigh points that might accumulateair or vapor leading to reduced flowor even cessation of flow from airbinding. And back at the suction ves-sel, the submergence over the vesseloutlet must be sufficient to preventvortexing (Figure 9) or the outletmust have an effective vortex break-er.

The second factor to beaddressed while at the site is thepump's service history. This can beobtained from the plants' mainte-nance department. What needs to belooked for is evidence of problemsin the pump's application, materialsof construction or mechanicaldesign.

Evidence of poor applicationmight be indicated by frequent shaftseal and bearing failures, rapid wearat the running clearances, frequentshaft failure, noise and vibration or

➞➞

PUMP TOTAL HEAD

HEAD

SYSTEMHEAD

TESTFLOW

PRESSURE➞

➞➞➞➞➞

VALVE

PIPING FRICTIONTOTAL STATIC HEAD

ELEVATION

Figure 4. System head from measurements

a) Eccentric Reducer Flat Side Up b) Eccentric, Flat Side Down

Figure 6a & b. Correct installation of reducers

HEAD

TOTAL SUCTION HEADPRESSURE & STATIC

VAPOR PRESSURE

FLOWTEST

NET SUCTION HEAD

NPSHA➞

➞➞

➞➞FRICTION

Figure 5. NPSHA characteristic from measurements

Page 186: Centrifugal Pumps Handbook

The Pump Handbook Series 187

premature impeller erosion. Theseare all symptoms of prolonged oper-ation at low flows. Whether this isthe case can be determined by com-paring known flow rates with thepump's performance curve to seewhere it has been operating relativeto the pump's best efficiency point(BEP).

Improved construction materi-als are warranted if the pump has ahistory of component failure due togeneral corrosion, corrosion-erosion,erosion, fatigue or erosion-fatigue. Itis often difficult to differentiateamong these causes of componentfailure, so it may be necessary toconsult a metallurgist. In some caseschanging materials may not beenough. It may be necessary eitherto correct a problem in the process,such as lowering the concentrationof abrasive solids or bringing the pHcloser to neutral, or to change to amore suitable type of pump.

Mechanical design is suspectonly when the influences of applica-tion and the pumped liquid have

been eliminated. (This may be thereverse of common practice, but isthe sequence to be followed in trou-bleshooting pumps.) Strain causedby piping loads is a major cause ofmechanical problems. If the pumphas had a high incidence of seal,bearing, coupling or shaft failures,the cause may be piping loads. Thequestion then is whether the pipingloads are too high or the pump notstiff enough. A computer analysis ofthe as-built piping is the first step inresolving this question. If the pipingloads are reasonable or high butcan't be changed, a switch to a pumpwith higher piping load capacitymay be necessary.

Short MTBR caused solely bypump mechanical design is rare inmodern designs, but not uncommonin designs dating back 30 years ormore. The usual difficulties are rotorstiffness, rotor construction, bearingcapacity, bearing cooling, bearinghousing stiffness and casing andbaseplate stiffness. These problemstypically manifest themselves as fre-quent seal, bearing and shaft failuresand rapid running clearance wear.Most of these are also symptoms ofpoor application, so care is needed insorting out the true cause of theproblem.

PUMP OPTIONS FOR UNIT REVAMPArmed with accurate data on the

system head and NPSH available,and knowing whether the suctionpiping or pump need correction aspart of the revamp, it's time to lookat what has to be done and how bestto do it.

First, data developed by theprocess designer must be checkedagainst the actual system head andNPSH available characteristics andcorrected if necessary. As indicated,the questions to be answered at this

Application NPSH Margin% NPSHR3________________________________

Water, cold 10-35(1)(2)

Hydrocarbon 10 (2)

Boiler feed, small (3) 50

Table 1. Typical NPSH Margins

S➞➞➞

14-12-10-

8-6-4-2-0- ' ' ' ' '2 4 6 8 10

SS (

feet

)

V (ft/sec)

VV

Figure 9. Submergence over suction vessel outlet

Figure 8a & b. Elbows at the suction of double suction pumps

a) END SUCTIONSOURCE BELOW

b) END SUCTION >Ns – 3,200 (2,750) orS > 11,600 (10,000)SOURCE ABOVE

➞➞

>10D2➞

>4D1

ECCENTRIC

S

S

S

>10D2

>4D1

CONCENTRIC

S➞

Figure 7. Suction piping layouts with two elbows in series

a) Desirableb) Undesirable

suction

path of water

path of water

suction

long radius elbow

Page 187: Centrifugal Pumps Handbook

188 The Pump Handbook Series

point are the following. How muchpressure drop does the control valveneed to function reliably? Is it moreeconomical to change to a variablespeed pump? What NPSH margin isneeded to ensure rated pump perfor-mance and expected life? The first isa question for the valve designer.Table 1 provides a starting point forthe third.

Notes:1. depends on size, higher mar-

gin for larger pumps2. minimum 3 feet3. up to 2500 hp at 3600 rpm4. U1 greater than 100 fps

New service conditions for theunit revamp can be met by exercis-ing three options:

• rerate the existing pump orpumps

• buy an additional pump orpumps of the same design

• buy pumps of a new design

These choices may, in turn, beinfluenced by the operating history ofthe existing pumps. Table 2 summa-rizes the needs developed from inves-tigation of the existing pumps and theusual options for satisfying them. Eachof the options is then discussed, start-ing with hydraulic considerations.

Rerating the existing pump isthe simplest course. To be success-ful, the rerate must be designed tomeet the new conditions of serviceand at the same time overcome anydeficiencies in the original applica-tion, such as being oversized for thenormal flow or having a suction spe-cific speed that is too high.

Adding a pump, either in paral-lel or series, is one way to achievesubstantially higher flow or head,respectively. To do so successfullyrequires care. For pumps in parallelthe fundamental rule is that thepumps operate at the same head;

therefore, the combined head capac-ity characteristic is developed byadding capacities at the same head(Figure 10).

Two cautionary comments areneeded here. First, to share flow reli-ably, the head of each pump mustrise continuously to shutoff, or to theminimum bypass flow in the case ofmultistage pumps. Second, theincrease in flow with each additionalpump depends on the steepness ofthe system head characteristic (Fig-ure 10). When the system headcurve is steep, as with referencecurve B in Figure 10, the increase intotal flow with two pumps is quitesmall. In this case the flow per pumpcan be well below BEP, with theresult that the MTBR of the pumps isreduced. At the other extreme, if therating of the pumps is also increasedby changing to a larger impeller, the

runout NPSHR and power of a singlepump must be checked to ensurethat the pump has enough NPSHmargin and that the driver will notbe overloaded (Figure 10).

Pumps in series operate at thesame capacity unless flow is takenfrom between them. Their combinedhead capacity characteristic is there-fore developed by adding heads at thesame capacity (Figure 11). Usingseries pumps against a system withhigh static head (curve A in Figure 11)poses the risk of flow cessation if oneof the pumps shuts down. This possi-bility must be taken into account incalculating the degree of redundancybuilt into the system and in designingthe pump control system.

Beyond this hydraulic considera-tion are two mechanical issues relat-ing to pressure containment. First, thecasing of the second pump must havea maximum allowable working pres-sure (MAWP) greater than the maxi-mum discharge pressure developedwhen both pumps are running atshutoff with maximum suction pres-sure. Second, the shaft seal(s) of thesecond pump will, in most designs, besealing suction pressure. Unless thepumps are separated in elevation, thesuction pressure of the second pumpis close to the discharge pressure ofthe first. This must be recognized inselecting the shaft seal.

Buying new pumps is the mostcomplex option, but it is necessarywhen the new service conditions arebeyond the capability of the existingdesign.

Regarding mechanics, a hydraulicrerate of older pump designs typicallyis combined with a mechanicalupgrade to raise MTBR. Most manu-facturers now have standard upgradesavailable for pumps ranging fromsingle stage overhung to multistage.

If the existing pumps have suf-fered corrosion or erosion abnormalfor the service and the class ofpump, a change of materials shouldbe considered. On the other hand, ifthe corrosion or erosion appearsmore related to the type of pumpthan to the service, changing thepump might be more economical inthe long run.

In rare circumstances it will beclear from an existing pumps' ser-

NEED OPTIONSRerate Add Replace Materials Construction

Lower flow ∆Higher flow - small ∆Higher flow - large ∆ ∆Corrosion resistance ∆ ∆Erosion resistance ∆ ∆ ∆Better mech design ∆ ∆

H.

'B'

B1 A2B2

FLOW

2 PUMPS

1 PUMP

'A'

Figure 11. Pumps in series

1 PUMPPOWER

NPSHR3

RUNOUT FLOW

2 PUMPS

A. FLAT SYSTEM HEAD

B. STEEP SYSTEM HEAD

A

B

AB

H.

Figure 10. Pumps in parallel

Page 188: Centrifugal Pumps Handbook

The Pump Handbook Series 189

vice history that it is the wrong con-figuration for the service. Reratingsuch a pump to an even higher ener-gy level would simply aggravate thiscondition. A high energy overhungpump in a severe service is a typicalexample. So as not to jeopardize thesuccess of the entire project, a pumpthat is clearly not suitable for its ser-vice must be replaced.

Once an option is selected, thenext step is to prepare the specifica-tion. The essential rule for a goodspecification is to keep it simple.Many a good solution has turnedinto a purchasing nightmare to thedetriment of the revamp projectbecause those preparing the pumpspecifications forgot this simple rule.The elements of a simple specifica-tion are:

• a one or two page data sheet• scope of supply summary, sup-

plemented by a terminal pointdiagram if needed

• agreed upon terms and condi-tions

A more detailed discussion ofthis phase and the equipment pur-chasing options can be found in thearticle "Pump Buying Strategies" byJ.T. McGuire in the January, 1993issue of Pumps and Systems. (Avail-able as part of The Pump HandbookSeries from the publishers of Pumpsand Systems.) ■

J.T. ("Terry") McGuire is directorof marketing for the Huntington ParkOperations Division of the Ingersoll-Dresser Pump Company.

Page 189: Centrifugal Pumps Handbook

190 The Pump Handbook Series

Pumping Hot Stuff –Another Perspective

Here are some general considerations for “hot” applications and, specifically, heat transfer fluid services.

By Pumps and Systems Staff

Many words can mean dif-ferent things to differentpeople – and equipmentmanufacturers. This truth

is important to appreciate whenselecting pumps for your applica-tion. Probably because manufactur-ers are often “industry oriented,”this situation seems to occur regular-ly in the case of high temperatureservices. For example, some pumpsuppliers serve the chemical processindustry (CPI) primarily while oth-ers sell mainly to petroleum refin-ers. In the former case, ANSIstandard pumps (Photo 1) are thenorm while API (American Petrole-um Institute, Photo 2) standards arethe design of choice in the latter. PerASME B73, the specification applic-able to ANSI models, pumps shouldbe designed to handle temperaturesup to 500º F. (Some manufacturershave provisions for higher limits,though non-metallic materials maybe limited to relatively low tempera-tures in the 200º F range.) APIpumps are available for much high-er temperatures. A number of ven-dors make hybrid or specialtyequipment that does not fit eitherstandard but is appropriate for hightemperature services.

What are the primary differ-ences between the two broad cate-gories of pumps relative to handlinghigh temperature liquids? The mostapparent is the pump mountingmethod. ANSI pumps are footmounted. API units are pedestal, orcenterline mounted. This means thepump is effectively mounted at thecenter-line of the shaft, allowing for

time. On the other hand, steel is the“lowest grade” casing materialoffered by API pump manufactur-ers, and 300 lb raised-faced flangesare also typical. In any applicationwhere the risk of fire is significant,such as a refinery service, the effectof thermal shocking of the casingmaterial during fire fighting effortsmust be taken into account. The lastthing petroleum processors wouldwant is to provide additional fuel toa fire as a result of a cracked ironcasing. API pumps also have moreextensive cooling options – even themounting pedestals are sometimescooled.

Another major considerationwhen selecting pumps for high tem-perature services is the sealing sys-tem. Packing may be used for someapplications, like boiler feed – inwhich case a smothering glandwould be employed. In the case ofmechanical seals, a film of liquid isneeded to cool and lubricate the sealfaces. While the high temperatureliquid being pumped may be appro-priate for this purpose, often it is notsuitable due to lack of lubricity. Thisnecessitates an external seal flushingarrangement. Consideration must begiven when selecting the fluid usedto avoid contamination of theprocess liquid. A combination flushand quench gland might also beneeded. The quenching functioninvolves an externally supplied flu-id, such as steam, to contain thepumpage that passes the seal faces.If it is important that the pumpedliquid not reach the atmosphere,either a multiple seal arrangement –

CENTRIFUGAL PUMPS

HANDBOOK

thermal growth of the pump in hightemperature services withoutimpacting the alignment with the dri-ver. A foot mounted pump has onlyone way to grow – up from the base.The result will often be misalignmentto the degree that bearing life will becompromised. Differences in materi-al and flange standards between theproduct types are also evident. ANSIpumps are commonly available withductile iron, stainless or other alloycasings and 150lb flat-faced flanges.Steel is sometimes an alternative, butit is rarely available with a short lead-

Photo 1. Example of a horizontalmetal ANSI process pump

(PHO

TOCO

URTE

SYOF

GOUL

DSPU

MPS

)

Photo 2. Example of an API processpump

(PHO

TOCO

URTE

SYOF

INGE

RSOL

LDR

ESSE

RPU

MPS

)

Page 190: Centrifugal Pumps Handbook

The Pump Handbook Series 191

possibly with barrier fluid system –or a sealless pump may be required.This latter category includes cannedmotor and magnetically drivenunits.

We do not have space here forexhaustive explanations of varioustypes of high temperature services.By exploring one of the most criticalpumping applications in processplants, however, we can illustratethe importance of these considera-tions in the handling of hot liquids.That category of pump servicesinvolves heat transfer.

If process temperatures are notmaintained (often within very nar-row ranges), the result can includeplant shut-downs and/or substantiallosses in product and production.Restarting the process can be anextremely complicated and time-consuming procedure. The typicalmethod to keep those temperatureswhere plant operators want theminvolves pumping one of severalheat transfer fluids (HTF’s) in “hotoil” recirulation systems. These flu-ids have a common trait – they areexpensive. Also, many are noxiousand even toxic to plant personnel –or potentially detrimental to theenvironment.

Looking at the options availableto a buyer of pumps for an HTF ser-vice will serve to illustrate some ofthe complicated issues facing thedecision maker. Several factorsrelated to costs and operation willneed to be evaluated. For example,if a mechanically sealed pump isselected, the pumped fluid will like-

ly prove to be anexcellent seal flushbecause of its typi-cally high lubricityand flash point,although it willprobably need to becooled before beinginjected at the sealfaces to assure rea-sonable seal life.This is usually doneby taking a smallamount of thepumpage off the dis-charge and piping itthrough an appro-priate cooler. Thesmall amount of flu-id that passes theseal faces will evap-orate into the atmos-phere. Without theuse of a quench, for-mation of crystals atthe seal faces can bea problem resultingin shortened seal life.Over time some make-up amountwill be needed. In the case of a sealfailure, obviously there could be asignificant amount of HTF lost, andarrangements to handle and replacethis material need to be made. Apopular mechanically sealed pumpused for HTF service in recent yearsis the Dean model R400 (photo 3.)This design has many API pump fea-tures, such as centerline mountingand bearing frame cooling provi-sions, but it is priced between ANSIdesigns and those that fully complywith API-610. A number of manu-facturers make pumps of this lattertype. If applicable, corporate policyor government regulations relativeto these fugitive emissions maynecessitate provisions for monitor-ing and reporting, as well as specificcontainment and processing proce-dures for pumpage that leaks.

Alternatively, and likely be-cause of these regulatory controlsand corporate mandates, many buy-ers are opting for one of the two gen-eral categories of sealless pumps.Though canned motor and magneti-cally driven designs typically have asomewhat higher first cost, theyoffer operational and, in some cases,

coincident cost advantages. Thecanned motor arrangement is, ineffect, an induction motor with itsrotor serving as the pump’s shaft. Athin metal can separates thepumpage from the stator. The rotorbearings are journal type and lubri-cated by the pumpage. In the case ofHTF, as for the mechanically sealedpump design, a small portion of thepumpage is often cooled – typicallythrough an integrally mounted cool-er, but one that in this case is used tocool the motor windings. A numberof manufacturers make cannedmotor pumps suitable for heat trans-fer service. Magnetically drivenpumps fall into one of two sub-cate-gories: synchronous and eddy-cur-rent. The first group utilizes one oftwo rare earth materials for both itsexternal and internal magnet assem-blies. Neodymium is significantlyaffected by heat, especially in therange of heat transfer fluid services,making this material unsuitable inhigh temperature applications. Moreappropriate in this temperaturerange is Samarium Cobalt, which,while the more expensive of the twomaterials, is impacted by high tem-perature operation to a significantly

HIGH TEMPERATURE EDDY CURRENT DRIVE

TORQUE RING

ROTATING OUTER MAGNET ASSEMBLY

CONTAINMENT SHELL

Outer Magnet Ring is rotated by the motor. TheMagnetic Flux which passes from one pole to the next,passes through the Torque Ring. This creates a flow ofEDDY CURRENTS. This power pulls the Torque Ring afterthe Outer Magnet Ring at a slightly slower speed whichis proportional to the power requirement.

Figure 1. Kontro, torque ring design

Photo 3. The Dean R400A mechani -cally sealed pump is used for HTFservice

(PHO

TOCO

URTE

SYOF

DEAN

MET

PRO)

Page 191: Centrifugal Pumps Handbook

192 The Pump Handbook Series

lesser degree. Several manufacturersserve the HTF market with appropri-ate units. The second group – magdrive pumps – isn’t really a group.The Kontro product line manufac-tured by Sundstrand Fluid Handlingoffers a unique torque ring design(Figure 1) that is unaffected by heatin the range for this application. Asin the typical mag-drive pump, per-manent outer magnets are coupled toa driver. Operating outside the fluid,they see reduced temperatures andare not affected by the heat as similarmagnets would be in an internalassembly. What makes this designdifferent is that power is transferrednot to a set of inner magnets but to arotating element consisting of theimpeller, shaft and integrally mount-

ed torque ring (with a coil much likethat in an electric motor.) This assem-bly runs in product lubricated journalbearings.

In all sealless options the elimina-tion of noxious emissions and potentialleaks from seal failures simplifies ornegates the need for internal corporateor external regulatory reporting, aswell as provisions for containment andsubsequent handling of fugitive emis-sions. The improved environment foroperators is often cited as a significantadvantage of using sealless pumps.Make-up requirements of expensiveHTF lost through seal operation andoccasional seal failures are eliminated.Interestingly, the somewhat lower effi-ciencies of sealless pumps actually con-tribute heat to the HTF, reducing the

load on hot-oil heaters. This is offset bycooling water requirements needed formany canned motor pumps, thoughmanufacturers do offer optional fea-tures, such as ceramic insulation, toallow higher temperature operationwithout cooling in lower horsepowersizes. Without the need for cooling ofthe recirculated lubrication fluid, a netheat gain is provided to the system inthe case of this latter group of cannedmotor pumps as well as the mag-driveand torque ring designs.

As you can see, there are many fac-tors to consider when selecting a pumpfor HTF or other high temperature ser-vices. First cost or electrical powerrequirements are two obvious consid-erations, but often the less apparentissues will have a significant bearing onthe purchase decision.■

Page 192: Centrifugal Pumps Handbook

The Pump Handbook Series 193

The Power of Speed and Staging

The energy consumption of most centrifugal pump applications can be reduced byoptimizing the specific speed. This can be done with either staging or shaft speed.

By Dave Carr and Bill Mabe

Centrifugal pump decisionsnaturally require purchasersto judge the suitability ofalternative selections. Pumps

comprise the second largest segmentof rotating equipment, and neverhave such decisions been more criti-cal. Intense competition dominatesnearly all of today’s markets. Theminimization of expenditures is aconstant focus with organizationsintent on survival. Energy costsassociated with pump drivers are aroutine target in the continuouseffort necessary to maintain marketviability. Clear recognition of themost efficient pump options to meetprocess “flow versus head” will armpump users with essential tools toattain reduced operating and main-tenance costs. Gains in centrifugalpumping efficiency can be readilyrealized through the optimization ofspeed and staging. We will present apractical approach to understandingthese issues using a simple formattabulated over a broad range offlows and heads.

Energy and EfficiencyAs with all laws of nature, the

key factor in minimizing the energyper unit time required to performwork on an on-going basis is the effi-ciency of the process. The same istrue with centrifugal pumps. In U.S.units, the power equation is {HP =H x Q x S.G./ 3960 / η} where HP ishorsepower, H is the total dynamichead of the pump in feet, Q is flowrate in gallons per minute, S.G. isthe specific gravity of the pumpedliquid (compared to water) and 3960

pump’s normal operating point andthe BEP, the less energy it will con-sume.

The second premise is that,inherent to the definition:

NS = N x Q0.5/H0.75 (1)

specific speed is directly propor-tional to rotating speed (N) in revo-lutions per minute and the squareroot of flow (Q) in gallons perminute while inversely proportionalto the three quarter power of head(H) in feet. Finally, there are physi-cal constraints to the best efficiencythat trend flow – that is, there isincreasing efficiency with increasingflow. This relationship is depicted inFigure 1 showing an expected bandof achievable efficiency within aflow range representative of the vastmajority of pump installations. Itwill be shown later that the effect offlow is more accurately related tothe size of the pump. This is animportant concept when consider-

CENTRIFUGAL PUMPS

HANDBOOK

is a constant. The pump’s efficiencyis usually represented by the Greeksymbol eta (η ) and is expressed as adecimal (i.e., 75% is written as 0.75).To lower consumed power effective-ly, the efficiency component is theonly variable factor in this equationfor a given set of process conditions.Options do exist when selecting orbeginning the design, however, sinceefficiency is a function of the pump’sspecific speed.

Specific Speed and Efficiency

Much has been written about“specific speed” (NS), a term thatrelates the pump performance toshape or type of impeller. It is notour intent to review the backgroundof the term, but a few basic premisesmust be explained. The first is that apump’s best efficiency point (BEP)corresponds to the head, flow, andspeed combination matching theimpeller’s design specific speed.Thus, the closer a match between a

10.90.80.70.60.50.40.30.20.1

0

50 100

200

500

600

800

1000

1200

1400

1500

FLOW GPM

Ns= 2000

500

EF

FIC

IEN

CY

Figure 1. Efficiency as a function of capacity and specific speed

Page 193: Centrifugal Pumps Handbook

194 The Pump Handbook Series

ing the potential efficiency to begained by increasing the specificspeed by raising the rotating speed.

Staging Reduces EnergyStaging is one way to increase

the efficiency for a given set ofhydraulics.

By dividing up the head by thenumber of stages (Z), the effectivespecific speed is increased by Z3/4.As a result, the overall efficiencyapproximately increases accordingto the relationship shown in Figure1. When referring to this figure, usestage head to calculate the specificspeed.

Speed Reduces Energy Specific speed is directly pro-

portional to the shaft rotating speedaccording to Equation 1. As thedesign speed is increased beyondtypical electric motor speeds, thespecific speed also increases for agiven head and flow. As discussedearlier, efficiency generally increas-es with an increase in specific speedand flowrate. One could concludethat it is always better to increasethe speed of a pump to maximizethe specific speed and the efficien-cy. As we shall see later, however,higher speeds may not always bethe best solution. Futhermore, mostpublished data, such as that shownin Figure 1 taken from Karrasik(1985), can be seriously misleadingfor high speed pumps.

High speed pumps are general-ly smaller than low speed singlestage or multistage pumps. Euler’sequation shows us that head rise isproportional to impeller tip speedfor a given exit blade angle. Tipspeed is the product of the impellertip diameter and the shaft rotatingspeed. Once the designer fixes theexit blade angle, the impeller diam-eter for a fixed design head decreas-es with increasing speed. Figure 2,established from experience with alarge number of high speed rocketengine turbopumps, shows that theefficiency of a small pump is consis-tently lower than the efficiency of alarger pump. Consequently, it is notaccurate to use Figure 1 directly forhigh speed pumps without correct-

ing the data for size effects.The effect of impeller size on

hydraulic efficiency at a fixed spe-cific speed is due to surface finishes,friction factors, vane blockages, andinternal leakage clearances that cannot be practically scaled down froma large size. As a result, the percent-age of hydraulic losses for smallpumps is significantly larger thanthat for larger pumps. For impellerdiameters greater than 10 inches,the size effect is small and generallyinsignificant. For diameters lessthan four inches, the efficiencypenalty is particularly severe. Theoptimum speed for a given head andflow is sometimes a tradeoffbetween the efficiency gain fromincreased specific speed and the effi-ciency reduction due to smallerimpellers. For most commercialpump applications, the gain fromincreased specific speed generallyoffsets the loss in efficiency due tosize effects at higher design speeds.

Speed or Staging?Tables of average hydraulic effi-

ciencies (Table 1 and Table 2) havebeen calculated over a range of typi-cal heads and flows for commercialapplications using data adaptedfrom Anderson (1980). The tablescan be used to compare efficiency atdifferent speeds and numbers ofstages. These efficiencies are forcommon motor speeds of 1800 and3600 rpm. The tables also includedata for shaft speeds of 7200 and14,400 rpm routinely obtained with

commercially available speedincreasing gearboxes. The calcula-tions include size effects.

Note that increasing flowratesand shaft speeds both tend to pro-duce higher pump efficiencies for agiven head rise. At a given flowrate,the efficiency generally increaseswith increasing shaft speed. Itshould also be obvious from the tab-ulated values that reducing the headper stage increases the efficiency.Efficiencies tend to maximize at thehigher flowrates with speed andmultiple stages having minimaleffects. The shaded portions of thetable are usually impractical selec-tions that one should avoid. Forthese applications, choose eithermultistage pumps, higher speedpumps, or mixed/axial flow impellerdesigns.

Overall Power The most prevalent pump dri-

vers in use today are two or fourpole AC electric motors. Speed inexcess of those capabilities, howev-er, requires the use of alternativeequipment. Turbines are routinelyused in process plants that have lowto medium pressure steam to attainspeeds as high as 10,000 rpm.Mechanical gearboxes, both integraland free standing designs, are thenorm with electric motors, but beltand sheave arrangements can alsobe used. Speed ratios as high as 10:1are possible with some gearboxesgiving the pump designer more thanenough flexibility to optimize specif-

10.90.80.70.60.50.40.30.20.1

0

500

600

700

800

900

1000

1100

1200

1300

1400

1500

1600

1700

1800

1900

2000

SPECIFIC SPEED

EF

FIC

IEN

CY

10

4

DIAMETER, IN.

1

Figure 2. The effect of size on efficiency

Page 194: Centrifugal Pumps Handbook

The Pump Handbook Series

195

5075

100125150175200225250275300325350375400425450475500525550575600625650675700800900

100012501500

HEAD PER STAGE, FT HEAD PER STAGE, FT HEAD PER STAGE, FT250

SPEED, RPM:500 750

SPEED, RPM:

FLOW, GPM

SPEED, RPM: SPEED, RPM:

1800 3600 7200 14,400 1800 3600 7200 14,400 1800 3600 7200 14,400

0.550.610.640.670.690.700.720.730.730.740.750.760.760.770.770.770.780.780.790.790.790.790.800.800.800.800.800.810.820.820.830.83

0.620.660.690.710.720.730.740.750.750.760.760.770.770.770.770.780.780.780.780.780.780.780.790.790.790.790.790.790.790.790.790.79

0.550.600.640.660.680.690.710.720.730.730.740.750.750.760.760.760.770.770.770.780.780.780.780.780.790.790.790.800.800.800.810.81

0.490.550.590.620.640.660.670.690.700.700.710.720.730.730.740.740.750.750.750.760.760.760.770.770.770.770.780.780.790.800.810.81

0.570.610.640.670.680.690.700.710.720.730.730.740.740.740.750.750.750.750.760.760.760.760.760.760.770.770.770.770.770.780.780.78

TA

BLE

1 C

EN

TR

IFU

GA

L PU

MP

HY

DR

AU

LIC E

FF

ICIE

NC

Y

0.440.510.550.580.610.620.640.660.670.680.690.700.700.710.720.720.730.730.740.740.750.750.750.760.760.760.770.780.780.790.810.82

0.360.430.480.520.540.570.580.600.610.630.640.650.660.660.670.680.680.690.700.700.710.710.710.720.720.730.730.740.750.760.780.79

0.270.350.410.440.470.500.520.540.550.570.580.590.600.610.620.630.630.640.650.650.660.660.670.670.680.680.690.700.710.720.750.76

0.630.660.680.690.700.710.710.710.720.720.720.720.720.720.720.720.720.720.720.720.720.720.720.720.720.720.720.710.710.710.700.69

0.430.490.540.570.600.620.630.650.660.670.680.690.700.710.710.720.720.730.730.740.740.750.750.750.760.760.760.770.780.790.810.82

0.600.640.670.680.700.710.720.720.730.730.740.740.740.740.750.750.750.750.750.750.750.760.760.760.760.760.760.760.760.760.760.76

0.170.260.310.360.390.420.440.460.480.490.510.520.530.540.550.560.570.570.580.590.600.600.610.610.620.620.630.640.660.670.700.72

NOTE: 250 > Ns > 4000 AND DIAMETER > 15 IN.

SPEED, RPM: SPEED, RPM:

Page 195: Centrifugal Pumps Handbook

5075

100125150175200225250275300325350375400425450475500525550575600625650675700800900

100012501500

HEAD PER STAGE, FT HEAD PER STAGE, FT HEAD PER STAGE, FT1000 1250 1500

SPEED, RPM:

FLOW, GPM

SPEED, RPM: SPEED, RPM:

1800 3600 7200 14,400 1800 3600 7200 14,400 1800 3600 7200 14,400

0.450.510.550.580.610.630.640.660.670.680.690.700.700.710.720.720.730.730.730.740.740.750.750.750.760.760.760.770.780.790.800.81

0.540.590.620.650.660.680.690.700.710.720.720.730.730.740.740.740.750.750.750.760.760.760.760.760.770.770.770.770.780.780.790.79

0.510.570.600.630.650.660.680.690.700.700.710.720.720.730.730.740.740.740.750.750.750.760.760.760.760.760.770.770.780.780.790.79

0.480.540.580.610.630.650.660.670.680.690.700.710.710.720.720.730.730.740.740.740.750.750.750.760.760.760.760.770.770.780.790.79

NOTE: 250 > Ns > 4000 AND DIAMETER > 15 IN.

TA

BLE

2 C

EN

TR

IFU

GA

L PU

MP

HY

DR

AU

LIC E

FF

ICIE

NC

Y

0.030.120.180.230.270.300.320.350.370.380.400.410.430.440.450.460.470.480.490.490.500.510.510.520.530.530.540.560.570.590.620.64

0.240.330.380.420.450.480.500.520.530.550.560.570.580.590.600.610.620.620.630.640.640.650.650.660.660.670.670.680.700.710.730.75

0.090.180.240.290.320.350.380.400.420.430.450.460.470.480.500.500.510.520.530.540.540.550.560.560.570.570.580.600.610.630.660.68

0.300.370.430.460.490.520.540.560.570.580.600.610.620.620.630.640.650.650.660.670.670.680.680.690.690.690.700.710.720.730.750.77

0.410.470.520.550.580.600.620.630.640.650.660.670.680.690.690.700.710.710.720.720.730.730.730.740.740.740.750.760.770.770.790.80

0.000.070.130.180.220.250.280.300.320.340.360.370.390.400.410.420.430.440.450.460.460.470.480.480.490.500.500.520.540.560.590.61

0.200.280.340.380.410.440.460.480.500.510.530.540.550.560.570.580.590.590.600.610.610.620.630.630.640.640.640.660.670.690.710.73

0.370.440.490.530.550.570.590.610.620.630.640.650.660.670.680.680.690.690.700.700.710.710.720.720.720.730.730.740.750.760.780.79

SPEED, RPM: SPEED, RPM: SPEED, RPM:

196The Pum

p Handbook Series

Page 196: Centrifugal Pumps Handbook

The Pump Handbook Series 197

ic speed and, thereby, pump effi-ciency. Mechanical efficiencies gen-erally decrease somewhat withincreasing speed ratios. Regardlessof which approach is taken, the effi-ciency of the drive train must be fac-tored into the pump evaluation toensure that maximum overall effi-ciency is attained.

How to Use the TablesUse the tables to determine the

potential for improving the efficien-cy of a pump for a given head andflowrate. Both the effects of stagingand higher speed can be evaluatedquickly. The following exampleswill illustrate:

For head, flow, or speeds notgiven explicitly in the table, use lin-ear interpolation. Accuracy is suffi-cient for most comparison purposes.

Concluding RemarksToday’s marketplace presents

ample opportunity for pump usersto conserve energy through the opti-mization of pump efficiency. Perti-nent information regarding the useof speed and staging can assist withan educated pump selection. On thesurface, it may appear that allincreases in rotating speed or a high-er number of stages will result incomparable increases in efficiency,as a result of higher specific speeds.Remember, however, that the opti-mum speed or number of stages fora given set of hydraulics may be atradeoff between efficiency gainsand losses as a result of specificspeed and size effects. A disciplineduse of the accompanying table willgive the pump user an appreciationof these tradeoffs, particularly asthey impact the search for minimum

consumed power.■

Dave Carr is a Project Manager withSundstrand Fluid Handling (Arvada,CO). A graduate of Purdue Universitywith a B.S. degree in MechanicalEngineering Technology and M.S. inManagement, he has held variouspositions within the company’s sales,marketing, product engineering andproduct development departments.Bill Mabe has been an EngineeringManager with Sundstrand for morethan 20 years and is primarily respon-sible for new product development. Agraduate of the University of Missouriat Rolla with a Mechanical Engi-neer-ing degree, he is a member of theadvisory committees for Pumps andSystems magazine and the Interna-tional Pump Users Symposium.

REFERENCESKarrasik, I. et. al. (1985), Pump

Handbook, pp. 2,13, McGraw Hill:New York

Furst, R. B. et. al. (1973), LiquidRocket Engine Centrifugal Flow Tur-bopumps, NASA Space VehicleDesign Criteria (Chemical Propul-sion) SP-8109, pp. 15-17.

Anderson, H. H. (1980), Predic-tion of Head, Quantity, and Effi-ciency in Pumps, PerformancePrediction of Centrifugal Pumps andCompressors. ASME, pp. 201-211.

Example 2.

Head = 750 ftFlow = 100 gpm

Single stage efficiency:At 3600 rpm = 48%At 7200 rpm = 59%At 14,400 rpm = 64%

Note the dramatic increase in effi-ciency by choosing a higher speedpump for this typical single stageapplication.

Example 3.

Head = 1000 ftFlow = 175 gpm

At 3600 rpm:Single stage efficiency = 52%Two stage efficiency = 55%Four stage efficiency = 70%

Note the increase in efficiencyavailable with a four stage pump.Alternatively, one can select a sin-gle stage pump at higher speed.

At 14,400 rpm, efficiency = 68%

Example 1.

Head = 500 ftFlow = 300 gpm

At 3600 rpm:Single stage efficiency = 69%Two stages (250 ft/stage) = 75%

Or, at 7200 rpm:Single stage efficiency = 74%

Specify either increased speed orstaging to improve efficiencyeffectively.

Page 197: Centrifugal Pumps Handbook

Over the past decade a num-ber of significant articles inthe trade press have focusedon the wide variety of avail-

able pump designs that affect flowcharacteristics, service life, mainte-nance, safety, product purity, oper-ating costs and other factors. Toooften these erudite articles – writtenby highly specialized hydraulic,mechanical or materials engineers –are more valuable to pump designersand manufacturers than to thosewho specify, purchase, operate ormaintain the equipment. This obser-vation is particularly valid in the areaof corrosive, abrasive, hazardous orultra-pure fluids.

The adage that necessity is themother of invention is a good expla-nation for the extensive design andmaterial variations available.

The whole field of exotic metaland non-metallic pumps has beenand continues to be market driven.The standard metal pump market iscommodity oriented. In the pumpingof neutral fluids such as potablewater or ordinary wastes, the keyfactors are initial costs, availabilityand service. Design standardizationand available modifications areaimed at cost reduction and operator-friendly aspects. Those responsiblefor price, delivery and dependableservice are the main decision mak-ers.

All of these purchasing influ-ences are important when it comes topurchasing or specifying pumps forcritical service applications. Howev-er, the handling of acids, caustics,solvents, halogens, salts, contaminat-ed water, process fluids, ultrapurewater or reagent grade chemicals andpharmaceuticals demands insightinto design variations and materialchoices that are often specific to thefluid being handled or the operationinvolved. Pump design is still marketdriven but the speed with which newdesigns are translated into productavailability at acceptable costs is con-trolled by the size of the market forthe solution being offered. Plantengineering and maintenance per-sonnel play critical roles. So do con-sulting and system engineersplanning new installations or upgrad-ing existing ones.

Historical OverviewCenturies of pumping experi-

ence involving water and pH neutralor mildly corrosive liquids lie behindthe standard design of horizontalcentrifugal pumps. As new materialsof construction were developed,pumps became more compact, moreefficient and more resistant to mildlyacidic or alkaline solutions, as well asto various atmospheric conditionsand temperature variations. With thegrowth of the chemical processing

industries, pumps manufactured ofstainless steel (types 304 and 316)became industry standards, with themolybdenum bearing type 316 hold-ing sway because of its broader resis-tance to corrosion. As the marketdemanded even greater resistance,particularly to sulfuric acid, Alloy 20became prominent.

For even greater resistance toheat, corrosion or abrasion, highernickel alloys and various exoticmaterials were made available. Costswere driven up as alloy content rose.

Plastic pumps came into beingabout 50 years ago when heart/lungoperations made it necessary to han-dle blood without contaminating it ordestroying fragile red blood cells.The flexible liner, peristaltic typerotary pump developed for this life-saving critical application utilized anacrylic material called Lucite for thecasing or pump body and a pure gumrubber as the flexible component.These were the only two parts incontact with the fluid. The gentle“squeegee” action on the blood andthe use of noncontaminating non-metallic materials solved the poten-tial problem of red blood celldestruction and helped save lives.This unique sealless pump designfilled an unmet need for transferringcorrosive/ abrasive fluids at lowflows without corrosion, contamina-tion or seal leakage. Coincidentally,

A Guide to ANSICentrifugal Pump Design

and Material ChoicesHere’s a pragmatic look at the commercially available design variations and material

choices now being offered on pumps for corrosive/erosive and ultrapure service.

By Dan Besic

198 The Pump Handbook Series

CENTRIFUGAL PUMPS

HANDBOOK

Page 198: Centrifugal Pumps Handbook

there was a major investment byDuPont and others to develop anextensive line of synthetic elastomerssuitable for the liner. Industryresearchers also sought to providesolid chemically resistant plasticssuch as Teflon, reinforced fluo-ropolymers, polyethylene andpolypropylene for the casing. Themix and match potential for casingsand liners (a dozen of these are nowstandard) has metamorphosed this“invention” for a particular probleminto an almost universal design, a rel-atively low cost do-everything, seal-less plastic pump.

In another application, pumpusers handling bromine demandedimproved service life and betterworker environments than theywere getting from stainless steel ornickel pumps. The answer found bypump designers relied on a new fluo-ropolymer, trade named Kynar, forall structural components. The mar-ket driven need for this corrosionand abrasion resistant material hasso universalized its application that ithas since become a standard recom-mendation for centrifugal pumps inan extended list of difficult-to-handlechemical-related processes, as wellas those demanding high purity, suchas pharmaceuticals or chemicals.

More recently, the demand fornon-metallic ANSI horizontal cen-trifugals dimensionally interchange-able with metal ANSI pumps hasbecome sufficiently large to enablepumps made of polypropylene,polyvinyl chloride and to be cost-effectively manufactured for a widerange of processing operations, aswell as for water treatment andwastewater handling. For all theseand similar market related reasons,non-metallic pumps in a variety ofthermoplastic and thermoset formu-lations have emerged as serious alter-natives to metal ones. The burden ofreasonable choice, in many cases, isno longer linked to a particular inno-vative manufacturer of a specificpump in a specific material for a spe-cific application. Users now have achoice among many metals and non-metallics.

The purpose of this article is topresent an overview of commerciallyavailable design variations and mate-

rial choices now being offered tousers and specifiers of centrifugalpumps for corrosive/erosive andultrapure service. When in doubt,there is no substitute for direct inter-change with the manufacturer andyour own or their experience withthe specific fluids and service condi-tions involved with an application.

Centrifugal Pump DesignTaking the ANSI end suction

centrifugal process pump as the stan-dard, let us compare the design char-acteristics of the critical componentsmade with thermoset and thermo-plastic pumps, assuming for themoment that all these material choic-es are satisfactory for the application.First, a word about plastic-lined met-al pumps. Plastic-lined metal pipe isproving to be an economicalapproach to transferring corrosivefluids, but when it comes to pumps,linings are not as easy to apply ormaintain. They are, however, on themarket. Generally, pumps are pro-duced with a corrosion resistant lin-ing applied in the wetted areas. Mostcommon designs are those with fluo-ropolymer or thermoset liningsapproximately 1/8-3/16” thick. Thesepumps may be subject to early fail-ure from erosion, pin holing, peelingor physical damage from solid parti-cles, as well as from temperaturefluctuation. Resistance to corrosionand abrasion in a pump is a lot morecomplicated than it is in a pipeline.

Since plastic lined pumps do nottypically offer service life compara-ble to that achieved in pumps havingsolid molded plastic pump casingsand impellers, the design variationsof lined pumps will not be covered inthis article.

CasingDimensionally, all pumps meet-

ing ANSI B73.1 specification forprocess pumps are the same. Thechoice is yours. You can remove ametal pump and replace it with aplastic one, or vice versa, withoutchanging the piping. As the need fora non-metallic pump to replace thestainless steel pumps in corrosiveservice became evident, it was natur-al for the manufacturers to offerdesigns made of various fiberglass

reinforced plastics. The thermosetshave physical properties closer tometals than the thermoplastics, andthey combine this with many of theweight and chemical resistant advan-tages of non-metallics. Pressure rat-ings at ambient temperatures forflanges in thermoset pumps areequivalent to 150 lb. metal flanges ofthe same dimension. The pressure-temperature gradient is a linear fac-tor and starts to degrade the materialabove 100ºF, so if fluid temperaturesrun higher, some pump manufactur-ers will add metal backup rings tothe thermoset support head, or theyhave support heads which include abackup ring.

Thermoplastic pumps designedso that the flanges and casings arecompletely supported and protectedby heavy sectioned metal armor pre-sent no problem with respect to thepressure temperature gradient. Anyof the available armored ANSIthermoplastic pumps will conform tothe standards without modification.

Nozzle LoadingsAs a general rule, the allowable

nozzle loadings for thermoset pumpsare lower than those indicated formetal pumps. Published data suggestthat these pumps be exposed to noz-zle loads 1/2 to 2/3 that of similarlysized stainless steel pumps. Armoredthermoplastic pump designs con-forming to ANSI B73.1 specificationshave the discharge pipe and flangesreinforced (metal armored) so thatthey can accommodate the same noz-zle loadings as the steel pumps theyreplace. Connections to metal pipeput the full load on the metal armorrather than on unsupported plastic asis the case with thermoset designs.

ShaftsMost non-metallic pumps use

stainless steel or other stress proofsteel shafts. Regardless of whichmaterial is used, it is recommendedthat the shaft be completely sleevedin plastic to isolate it from the fluid inthe wetted area.

High strength carbon steel canbe safely used when the shaft is sosleeved. The sleeve can be made ofinjection molded glass reinforcedpolyphelene sulfide (PPS) when the

The Pump Handbook Series 199

Page 199: Centrifugal Pumps Handbook

200 The Pump Handbook Series

corrosive is relatively mild, but forbest results shaft assembly should bemade of the same material as thepump. Thermoplastic pumps shouldhave sleeves of PVC (Polyvinylchlo-ride), PP (Polypropylene), or PVDF(Polyvinylidene fluoride). The shaftsleeve in a horizontal centrifugalpump should be independent, notwelded to the impeller. When sodesigned, a damaged shaft sleeve canbe replaced by simply removing theimpeller and then the sleeve. Thisapproach is recommended becausemost users do not have plastic weld-ing capability.

Mechanical SealFor maximum corrosion resis-

tance it is important that mechanicalseal configurations be completelynon-metallic, or mounted so that theinboard or wetted non-metallic sealface is exposed to the fluid. Since sealchanges represent a significant main-tenance item, it is critical that thepump design permit easy inspectionof the seal without disassembling orremoving the impeller shaft. In stan-dard centrifugal pump designs whichare not made to ANSI specifications,or which do not permit back pull-out, the available work area for sealinspection and maintenance is cru-cial. In some ANSI, as well as non-ANSI pumps, a sliding bar design,which permits backing up the prima-ry seal for inspection without affect-ing shaft alignment, keeps downtimeat a minimum.

Externally mounted seals facili-tate maintenance of the pump/sealarea by allowing personnel to inspectseal placement visually. This permitsproper setting of the seals withoutrelying, with fingers crossed, onshims or measurement. Pump stuff-ing box designs, which permit thewidest choice of readily available sin-gle or double mechanical seals, pro-vide a definite advantage.

Shaft DeflectionOver the years the complicated

formula for shaft deflection, animportant component for comparingpotential seal life from centrifugalpumps, has been abbreviated L3/D4.This relationship between the lengthof the shaft overhang from the front

or inboard bearing to the impellerand the diameter of the shaft at theseal face is generally satisfactorywhen comparing two metal pumpsof similar design. However, it is use-less and misleading when comparinga metal pump with a non-metallicone. The reason is simple. The fullformula takes into consideration thediameter and the weight of theimpeller. Impeller diameters of engi-neered plastic pumps are equal to oronly slightly smaller than metal ones,but impeller weights are quite differ-ent. The lighter weight of the plasticimpeller and the correspondingreduced downward thrust or forceon the impeller end of the shaft areignored by the L3/D4 abbreviation.These variables significantly affectthe length/diameter ratio, making thecommon acceptable ratio (50) mean-ingless when applied to thermoset orthermoplastic pumps. Don’t getcaught in this trap. Insist on knowingthe actual vibration level at the bear-ings – the actual shaft deflection atthe seal face, not the “shaft stiffness”factor that the simplified formulaprovides.

Cost FactorsThere was a time when plastic

pumps were considered less expen-sive substitutions for metal ones.Those days are long gone. Non-metallic pumps for corrosive/ero-sive/hazardous or other specializedservices are carefully engineeredproducts with initial costs competi-tive to similar pumps of type 316stainless steel. Cost advantages, how-ever, become appreciable for thoseapplications involving the higheralloyed materials and for pumpsrequiring titanium, nickel and otherexotic metals. Published cost com-parisons indicate that plastic pumpstend to be 20% lower in price thanthose of Alloy 20 (a nickel-chromi-um-copper alloy originally developedfor handling sulfuric acid), and halfor less than the cost of Hastelloy C,the high nickel stainless alloy devel-oped for severe corrosive service.

Initial cost is but one factor,however. Service life is even moresignificant. Plastic pumps have beenin service sufficiently long in highlycorrosive/abrasive applications for

the purchaser or user to ask for dataon anticipated service life based onprevious installations. Laboratorytests on metal or plastic “coupons”prove to be poor guides to pump lifebecause they usually involve immer-sion time under static conditions.Corrosion or wear rates for partsrotating at 1800 or 3600 revolutionsper minute, or stationary parts han-dling fluid flows to 5000 gallons perminute, are not comparable to thosebased on laboratory tests. Ask thepump manufacturer for actual casehistory data to be sure.

There are other cost factors to beconsidered when comparing metalversus plastic pumps, or one plasticpump versus another. Metal to plas-tic weight comparisons impact oncosts involved in shipping, installa-tion, disassembly/assembly, andtime related problems such as gallingor seizing of threaded components.Seal life, spare parts requirementsand general maintenance require-ments must be considered. In all ofthese areas, the non-metallics offerclear cut advantages because of theirlighter weight (1/2 to 1/8 the weightof metal), lower component costs,non-corroding characteristics andlonger seal life.

The differential in cost betweenthermosets and thermoplastics isnegligible if we compare polypropy-lene and fiberglass reinforced plastic(FRP). Comparative service life ismore critical, but this has to be basedon the particular thermoplastic andspecific thermoset. Generally speak-ing, homogeneous thermoplasticshave much broader chemical resis-tance than thermosets, and ther-mosets offer physical propertieshigher than those of thermoplastics.For example, if the superior corro-sion and abrasion resistance of thefluoropolymers is required, the high-er cost of fluoropolymer thermoplas-tics such as PVDF or ECTFE is oftenrepaid many times over by theextended service life of and/or prod-uct purity advantages these fluo-ropolymers offer.

Maintenance FactorsStandard, routine preventive

maintenance programs are advisablefor all pumps regardless of materials

Page 200: Centrifugal Pumps Handbook

The Pump Handbook Series 201

of construction or design factors.Your own service experience is thebest guide to proper scheduling.

Where the installation is sub-stantially different from what youhave been involved with, it is wise torely on the specific experience of themanufacturer. When shifting frommetal pumps to plastic pumps, con-sider the following:

A. If you are connecting ther-moset horizontal centrifugal plasticpumps to metal pipes, you may needto use expansion joints to reduce noz-zle strain, particularly if there is asignificant difference in the modulusof elasticity of the two materials.This may not be necessary foramored.

B. For vertical centrifugalpumps used in deep sumps, makesure the design accommodates thedifferential in axial thermal expan-sion between the long plastic supportcolumns and the available impellerclearance. This, however, is not afactor for horizontal pumps.

C. Although non-metallic flan-ges are dimensionally the same asmetal ones, if they are unsupportedyou may need metal washers underthe bolt heads and nuts. Casing bolt-ing for non-metallics requires carefuladherence to torque specified by themanufacturer. This is true for allpumps but not as critical with metalor metal armored plastic ones.

D. Do not steam pressurizethermoplastic or thermoset pumps.This can cause components to bepressurized beyond ratings.

E. Protect outside mechanicalseal with a seal guard or cover.

Comparative MaterialCharacteristics

It is not the purpose of this arti-cle to concentrate on the long list ofsophisticated materials of construc-tion or erudite design factors overwhich the user has little or no con-trol. These are basically the responsi-bility of pump manufacturers. Pumpusers need not be metallurgists ormaterials engineers. Their concern isand should be with understandingthe relationship between availableconstruction materials and the ser-vice conditions faced. Pump design-ers have an endless variety of

materials from which the criticalwetted end components can beselected. In reality, however, usersare limited to choosing from among arelatively short list of commerciallyavailable materials. Exotic metal orcustomized composites may berequired for pumping problems thatcan’t be solved with materials thatcan currently be produced economi-cally. Limited production potentialoften makes it difficult to producepumps of these materials at reason-able cost and delivery schedules. Thepump materials listed below arereadily available for your considera-tion. Familiarity with the basiccharacteristics of these materials willhelp you select the most cost effec-tive material for your applications.

Materials of Construction

MetalsStainless Steel (type 316): A

general purpose austenitic chromium(18%), nickel(8%), molybdenum(3%) alloy providing broad resistanceto a long list of acids, caustics, sol-vents. Widely used for its atmospher-ic corrosion resistance and inproducts requiring sterilization. Notrecommended for sea water, brine,bromine or strong oxidizing acids.

Alloy 20: Originally developed to resist various concentrations ofsulfuric acid for which type 316stainless steel wasn’t suitable. Thisstainless alloy has higher nickel(28%) than chromium (20%) content,less molybdenum (2%), plus copper(3%). It offers good resistance todilute and strong, as well as mixedacids, sulfate and sulfites, sulfurousand phosphoric acids, chlorides andbrines.

Hastelloy “C”: With an increasein the nickel content to approximate-ly 50%, this nickel, chromium (22%),molybdenum (13%, tungsten (3%)alloy is recommended for uses inwhich the stainless steel and 20alloys fail. It offers much greaterresistance to oxidizing acids and acidmixtures, wet chlorine, chlorides,mineral acids, sea water and plat-ing/pickling solutions.

Thermosets

Fiberglass Reinforced Plastic(FRP): Although there is an endless

variety of thermoset formulations,many indicated by trade names, themost common standard formulationsconsist of vinyl ester and epoxy resinreinforced by glass fibers. Since thetwo components bring different val-ues to the composite – vinyl for cor-rosion resistance, epoxy for solventresistance – competitive formula-tions can be customized for a particu-lar service. Thermoset materialsgenerally offer broad resistance tomost acids, caustics, bleaches, seawater and brine. Special formula-tions are available for mild abrasiveservice. Thermoset pumps of FRPhave higher physical properties thanthose of thermoplastics and are avail-able for flows to 5000 gpm, twice theflow of current thermoplastic offer-ings.

Thermoplastics

Polyvinyl Chloride (PVC):Widely used in chemical processing,industrial plating, chilled water dis-tribution, deionized water lines,chemical drainage and irrigation sys-tems. Good physical properties andresistance to corrosion by acids, alka-lies, salt solutions and many otherchemicals. Not suitable for solventssuch as ketones, chlorinated hydro-carbons and aromatics.

Chlorinated Polyvinyl Chloride(CPVC): Its chemical resistance issimilar to, but slightly better than,PVC. It is also stronger. Excellent forhot corrosive liquids, hot and coldwater distribution and similar appli-cations.

Polypropylene (PP): A lightweight polyolefin chemically resis-tant to organic solvents as well asacids and alkalies. Generally not rec-ommended for contact with strongoxidizing acids, chlorinated hydro-carbons and aromatics. Widely usedin water and wastewater applicationsand for laboratory wastes where mix-tures of acids, bases and solventsmay be involved.

Polyvinylidene Fluoride (PVDF):Strong, tough and abrasion resistantfluoro-carbon material. Resists dis-tortion and retains most of itsstrength at elevated temperatures.Recommended for ultrapure waterand reagent chemicals. Resistant tomost acids, bases and organic sol-

Page 201: Centrifugal Pumps Handbook

202 The Pump Handbook Series

vents and equally suited for handlingwet or dry chlorine, bromine andother halogens.

Ethylene Chlorotrifluoroethyl-ene (ECTFE): Resists an extremelybroad range of acids, alkalies, organ-ic solvents and combinations ofthem, as well as other corrosive andabrasive liquids. Also resistant to oxi-dizing acids and hydroxides. Ideal forultrapure water applications.

Polytetrafluoroethylene(PTFE): This crystalline polymer isthe most inert compound known. Ithas useful mechanical properties atelevated temperatures. Impact resis-tance is high, but tensile strength,wear resistance, and creep resistanceare low in comparison with otherengineered plastics. Its coefficient offriction is lower than almost any oth-er material.

Selection CriteriaTemperature: Temperature

parameters are not critical in deter-mining the choices among the vari-ous metals. Metallurgically, thedifferences may be great, but all ofthe metals considered for handlingcorrosive fluids are stable at mostoperating temperatures. When itcomes to the plastics, however,anticipated temperature fluctuationsare very critical. Table 1 providesupper temperature limits for the non-metallic pump materials currentlyavailable.

Weight: The specific gravity ofthe material can be a significant vari-able when comparing metal to plasticpumps because of weight-relatedcosts such as shipping, installation,support structures and in-plant repo-sitioning. It is also vital in the com-parison of various plastics with eachother. As anyone who has lifted a flu-oroethylene polymer casing can tellyou, not all plastics tip the scale inthe same way. A PVDF casing may

weigh twice as much as one moldedin polypropylene. Of course, whencomparing metal-armored plasticpumps or plastic-lined pumps withall metal ones, the apparent differ-ences are minimized. Table 2 showsthe comparative weight of the mate-rials of construction. If the weightfactor is significant, ask the manufac-turer for total pump weights. Youmay need these to assist in installa-tion and piping.

Abrasion Resistance: Strange asit may seem, stainless steel pumpsare relatively poor compared to plas-tics when it comes to resisting wearfrom abrasive fluid streams. A major reason for this is that the oxidizedsurface which protects passivatedchromium-nickel stainless steel fromcorrosion is continuously removedby abrasive particles. The smooth,

PVC Polyvinyl chloride 140ºF

PP Polypropylene 185ºF

PE Polyethylene 200ºF

CPVC Chlorinated polyvinyl chloride 210ºF

FRP Fiberglass reinforced plastic 250ºF

PVDF Polyvinylidene fluoride 275ºF

ECTFE Ethylene chlorotrifluoroethylene 300ºF

PTFE Polytetrafluoroethylene 500ºF

Table 1. Temperature

Material Weight Loss/1000 Cycles

Polyethylene (UHMW) 5 mg.

PVDF 5-10

PVC 12-20

PP 15-20

CPVC 20

Stainless Steel 50

FRP 388-520

PTFE 500-1000

Table 3. Abrasion resistance (TaborAbrasion Tester)

Weight SpecificGravity

PP .91

PE .92 - .94

PVC 1.30

CPVC 1.49

PVDF 1.75

ECTFE 1.75

PTFE 2.14 - 2.20

FRP 3.4 - 5.0

316 Stainless 7.9

Table 2. Weight differentials of construction materials

Material Tensile Strength Hardness Impact(psi) R = Rockwell (IZOD)

D = Shore B = Barcal

PE 3,500 - 5,600 R 35-40 1.5 - 12.0PP 4,000 - 5,000 R 80-110 0.5 - 2.2PTFE 2,000 - 5,000 D 50-55 3.0PVDF 5,500 - 8,250 D 80 3.6 - 4.0ECTFE 6,500 - 7,500 D 75; R 93 No break at 73ºFPVC 6,000 - 7,500 R 113 0.4 - 2.0CPVC 7,500 - 11,000 R 121 0.6FRP 10,000 - 13,000 B 35-40 –

Table 4. Strength of the non-metallics

Page 202: Centrifugal Pumps Handbook

The Pump Handbook Series 203

uniform interior surface of a moldedthermoplastic casing is a significantfactor in reducing friction and turbu-lence, both of which contribute towear. The FRP materials rely on theepoxy/vinyl ester to contain theglass. If the surface is subject to fastflowing process streams containingsolid particles, abrasion can besevere. The reinforcing fibers may beexposed causing degradation of thecomposite or a wicking/bleedingaction that can contaminate the fluid.For these reasons, homogeneousthermoplastics offer much greaterresistance to abrasion. Table 3 showstypical weight loss of the materialsbeing considered.

Strength of Non-Metallics: Thedifferential in tensile strength andimpact resistance between metal andnon-metallic pumps may be of acade-mic interest, but insofar as users areconcerned, it is significant to keep inmind that unarmored non-metallicpumps need more care in handlingthan metal ones. Metal armoredthermoplastic pumps can be treatedas metal ones, but thermoset pumpsrequire a bit more attention. Rein-forced epoxy/vinyl composites canbe brittle (like cast iron), so precau-tions should be taken to provide pro-tection from falling overhead objects,fork lift trucks or careless handling.Tensile, hardness and impactstrength of the various non-metallicmaterials are shown in Table 4.

Corrosion/Chemical Resis-tance: When pump specifiers andusers are asked why they are consid-ering non-metallics, corrosion resis-tance is by far the number onereason given. The resistance of thebasic construction material is signifi-cant in terms of service life, safety,maintenance and, in many cases, thepurity of the product being pumped.This latter criterion is essential forthose using deionized water, reagentgrade chemicals or other ultrapurefluids in their manufacturing or pro-cessing operations. When metallic orother contaminants cannot be toler-ated, this resistance is criticalbecause it may seriously affect thequality or the value of the productbeing produced.

Many handbooks and materialsengineering articles give comparative

corrosion resistance data for anexhaustive list of metals and plastics.These are rated against page afterpage of different chemicals at vary-ing concentrations and at varioustemperatures. The tables are helpfulguidelines, but for the most part,they are based on corrosion or deteri-oration when the subject material isimmersed in the fluid under staticand unchanging conditions. Condi-tions of service in the real world areseldom that uniform or controllable.These lists and tables are no substi-tute for actual pumping experience –your own and that of your suppliers’.Described below are highlights ofsome interesting pumping operationsin which material selection played acritical role.

These examples may help yousee your own pumping problems in adifferent perspective. If this articleencourages you to review your com-pany’s pump purchasing and mainte-nance picture with a view to howyou might extend service life, reduceparts inventory, simplify mainte-nance procedures, improve productquality, meet current or anticipatedenvironmental regulations, or evenfeel comfortable about your currentoperating equipment and proce-dures, it will have served its purpose.

Examples:

1. Sulfuric Acid Waste StreamContaining 100 Micron Fines

This corrosive/erosive mixturewas destroying 316 stainless steelpumps. The short service life andcostly downtime could not be tolerat-ed. When the switch was made topolypropylene pumps to handle the 3 pH abrasive wastewater, the prob-lem was solved.

2. Corrosive/Abrasive MineWater

Acid run off with a pH of 1-2from coal mines in South Africa wereseverely corroding stainless steelpumps. The answer was found witha combination of non-metallics.Pump casings were supplied inpolypropylene, but a PVDF impellerwas specified because of the superiorabrasion resistance of this fluo-ropolymer. To reduce problems with

early seal failure, a Teflon (PTFE)packed gland and product flusharrangement was used instead of thestandard single seal product flush.The pump has been handling 16-18million gallons of acidic fluid perweek at 1800 gpm. No problems.

3. Etching Glass with Hydro-fluoric Acid

A major glass manufacturerexperienced severe pump mainte-nance problems when it came tohandling a slurry of hydrofluoricacid with a high content of propri-etary gritty compound. Originalequipment utilized a composite fiber-glass pump with a stainless steelimpeller. Service life averaged about2 months. The plant manager testeda variety of non-metallics and finallydecided on PVDF impellers for thecorro-sive/abrasive service andpolypropylene for the casing. To iso-late the shaft from the fluid, he spec-ified a thick sectioned PVDF sleeveand arranged to have the mechanicalseal reverse mounted. The pumpsare driven by 7.5 hp electric motorsat 1750 rpm. The 10% productionincrease experienced is credited tokeeping the production line operat-ing with thermoplastic pumps.

4. Acid Mixing Plant in Alaska

An oil field acidizing and wellservice company in Alaska ordered amodular plant built in Texas andshipped to their Prudhoe Bay site.Some of the pumps are polypropy-lene throughout. Others are fittedwith ECTFE impellers to handle thehydrochloric acid and zylene.Although the pumps are locatedindoors, the fluid temperature couldbe minus 60°F because the chemicalsare stored outdoors. Service reportsindicate no problems.

5. Circulating Phosphoric Acid

The use of lined metal pumps tohandle phosphoric acid was causingextensive shutdown due to pinholesand tears that allowed the metal cas-ings to contaminate the solution anddowngrade it. Part of the problemwas caused by heat generated in thepumps, which tended to destroy thethin linings. When solid PVDF ther-moplastic pumps were substituted,

Page 203: Centrifugal Pumps Handbook

product purity was assured and over-all maintenance substantiallyreduced.

6. Ultra Purity for HydrogenPeroxide Production

Specifications for these centrifu-gal pumps called for the use of pumpcasings, impellers and shaft sleevesto be made of virgin, unpigmentedfluoropolymer totally resistant to70% hydrogen peroxide beingpumped at 50gpm against a totaldynamic head of 80 feet. The materi-al selected was ECTFE, ethylenechlorotrifluoroethylene, which isnoted for its broad resistance tochemicals, pharmaceuticals and oxi-dizing acids. The seal rings werespecified in the fluoroelastomerViton and the casing jacket in TeflonPTFE.

7. Bleaching in White PaperMills

The most corrosive applicationsin white paper mills involve pump-ing chlorine bleach. Traditionally,this has been done with extremelyexpensive titanium pumps. Whenenvironmental concerns required aswitch from chlorine bleach to chlo-rine dioxide, pulp mills were able tochange from titanium to epoxy vinylester pumps at much lower initialcost, reduced spare parts inventoryand ready availability of pumps andparts.

8. Sulfuric Acid and a CausticChaser

A large can manufacturing divi-sion of one of the largest breweries inthe world faced costly downtime dueto excessive corrosion of the metalpumps transferring dilute sulfuricacid for the etch and strong causticfor the required neutralization. Thedecision was to change from stainlesssteel to PVC thermoplastic centrifu-gal pumps. All wetted end compo-nents were specified in the samehomogeneous thermoplastic. Theplastic pumps have served in thisround the clock operation for morethan a dozen years with only routinemaintenance.

9. Hot Oil/Hydrofluoric Acid/Caustic Mixture

A refinery using hydrofluoric

acid in a process to produce highquality gasoline ran into difficultpumping problems due to the vary-ing pH of the 150°F oil/HF/causticmixture. They tried a variety ofpumps from cast iron to 316 stainlesssteel, but a combination of corrosionplus impeller and seal damage fromaccumulated solids required pumpreplacement on a monthly schedule.Since installation of molded thermo-plastic polypropylene pumps withdouble mechanical seals and pressur-ized water jackets using 8 to 10 gphof clear water to cool the seal jacket,the pumps have performed flawless-ly.

10. Tall Oil Production in Kraft Mills

Tall oil soap, a byproduct of theKraft paper making process.Skimmed from the evaporated cook-ing liquor, the soap is acidulatedwith sulfuric acid in a reactor vesselto achieve a pH of 2.5 - 3.5 andallowed to settle out. The tall oil risesto the top, and the spend acid can beeither fresh acid (98%) diluted withwater to 30% or with spent acid fromthe bleaching operations. FRP pumpmanufacturers report that the ther-mosets show improved service lifeover pumps made of 316 stainless or20 Alloy.

11. Pumping Deionized Waterfor Laboratory Use

Medical laboratories utilizesealed diagnostic kits to analyzeblood samples. To assure correctdiagnosis, it is critical that the waterused to prepare the various chemicalsolutions be free of contaminants.Pretreatment of the processed waterrequires a sand filter, deionizationequipment and special fine pore fil-tration media to eliminate all partic-ulates. Once the water is purified, itis protected by a fluid handling sys-tem composed entirely of chemicallyinert plastics. It is stored in polyeth-ylene tanks and pumped through aclosed loop system of rigid PVC pipeby a horizontal centrifugal pumpwith all wetted components of PVDFfluoropolymer. This system usesclose coupled pumps with an integralpump/motor cantilevered shaft thatenables the pump to run dry for

extended periods without damage.

12. Pump Maintenance Reduced by $25,000

The extreme corrosiveness of anammonium chloride/zinc chloridesolution at 140°F resulted in the fail-ure of Alloy 20 pumps after a year ofservice. In addition to the high costof this annual pump replacement,mechanical seal failure occurred on amonthly basis requiring a minimumof 2 hours of lost process time eachmonth. When the metal pumps werereplaced with those made of virginthermoplastic polypropylene, the fol-lowing results were reported: lowerinitial pump cost, improved resis-tance to both corrosion and cavita-tion, an average seal life of ninemonths, and better than two years ofservice before removal of the pumpsfor reconditioning. Annual mainte-nance costs decreased by $25,000.

13. Silicon Wafer ProductionDemands PVDF Pumps

Production processes in themanufacture of high quality siliconwafers require the use of hydrochlo-ric acid, hydrofluoric acid and vari-ous caustic solutions. Themanufacturer had standardized onthermoset materials for the pumps,but the corrosiveness of the chemi-cals and mixed acid wastes provedtoo severe. Hazardous fluid leakageand reduced pump performance,particularly in the handling of thehydrofluoric acid waste streams,necessitated a changeover topolypropylene pumps with PVDFshaft sleeving in the wetted area.Plant management reports that leak-age and capacity problems are athing of the past.■

Dan Besic is Chief Engineer atVanton Pump & Equipment Corp. (Hill-side, NJ).

204 The Pump Handbook Series

Page 204: Centrifugal Pumps Handbook

W ith the advent of in-creasingly stringent en-vironmental regulationspertaining to storage tank

connections below liquid levels, self-priming centrifugal pumps are beingutilized in a wide variety of applica-tions. In the past, horizontal designshave been employed in these ser-vices due to the inherent advantagesof centrifugal pumps. Installationwas relatively simple, and if the suc-tion and discharge requirements of aparticular pump were met, satisfac-tory performance could be expected.

Converting a centrifugal pumpapplication that previously operatedoff of a positive suction head to oneconsisting of a negative, or combina-tion negative and positive suctionhead, requires additional criteria tobe met for a successful application.The addition of a compressible fluid(air) in the suction line imposes con-ditions on the piping system thatmust be overcome for a self-primingpump to remove air from the suctionline and pump fluid as a centrifugalunit is designed to do. Piping sys-tems composed of check valves, pipeloops and liquid traps should bereviewed with respect to the pumpoperation during the priming cycle toassure that all air can be evacuatedfrom the suction line prior to thepump moving liquid only.

Centrifugal Pump OperationThe basic theory of operation of

a centrifugal pump – from which itsname is derived – relates to centrifu-gal force. A rotating impeller impartsvelocity energy to the fluid betweenthe impeller vanes, and that velocity

is subsequently converted to pres-sure energy in the volute section ofthe pump casing.

This increase in pressure energy(total dynamic head on a pumpcurve) is proportional to the centrifu-gal force applied to the fluid. Sincecentrifugal force is directly propor-tional to weight, the differencebetween the discharge head of a cen-trifugal pump filled with water and acentrifugal pump filled with air is inthe order of 810 to 1! (This is the dif-ference between the density of waterand the density of air at atmosphericpressure.) A pump handling a mix-ture of air and liquid will exhibit dis-charge heads between these twolimits.

TDH ~ centrifugal force = wω2r g

w = weightω = angular velocityr = radiusg = gravitational acceleration

Density water at atmospheric pres-sure = 62.4#/ft3

Density air at atmospheric pressure= .077#/ft3

From this it is clear that largestatic heads cannot be imposed onthe discharge side of a centrifugalpump if it handles air during somephase of its operation.

Self-Priming CentrifugalPump Operation

There are many types of self-priming centrifugal pumps. Includedin this category are submerged

impeller designs such as vertical andsubmersible pumps and convertedhorizontal pumps utilizing auxiliaryvacuum producing equipment orsuction priming tanks that allow astandard centrifugal pump to operatewith a positive liquid head at alltimes.

The most popular self-primingcentrifugal pump used commerciallyis the “recirculation” or “peripheralpriming” type. It is characterized bya liquid reservoir either attached toor integrally constructed with the

pump casing (Photo 1). The suctionconnection is usually located abovethe impeller centerline in order tocontain and trap a volume of liquidused in the priming cycle.

Before the pump is initially start-ed, the liquid reservoir must be filledmanually. When the pump is shutdown, a syphon breaker or internalsuction check valve retains a quanti-ty of the pumped fluid in the reser-

Self-Priming Pumps: It’s in the System!Coordinate pump with piping system for optimum performance.

By Ray Petersen

The popular “recirculation” or“peripheral priming” type self-primingpump is characterized by a liquidreservoir either attached to or inte-grally constructed with the pump casing

(Pho

to co

urtes

y of F

ybro

c Pum

p Di

v.)

CENTRIFUGAL PUMPS

HANDBOOK

The Pump Handbook Series 205

Page 205: Centrifugal Pumps Handbook

voir for successive starts.The internal construction of a

self-priming pump is similar to aconventional centrifugal pumpexcept for the addition of a recircu-lating port in the volute passage thatis connected to the fluid reservoir(Figure 1). As the impeller rotatesduring the priming cycle, the liquidin the impeller and volute passage isdischarged out of the volute into anexpansion or air separation area ofthe liquid reservoir. However, beforethe air separation area of the pumpcan be filled with liquid and thepump considered primed and capa-ble of generating a pressure againstthe discharge piping system, air fromthe suction line enters the impellereye, causing a drop in pressure dueto the relative densities of air and liq-uid.

As the impeller continues torotate, the mixture of air and liquidmoving at high velocity will drawadditional liquid into the impeller

and volute area through the recircu-lation port. This mixture is then dis-charged past the volute cutwater intothe expansion area of the reservoir.In the expansion section, the liquidand air bubbles separate. The air,being lighter, vents upward out thepump discharge while the heavierliquid returns to the reservoir andcontinues to recirculate and entrainmore air. This cycle continues untilall the air in the suction line andimpeller is evacuated, at which pointthe pump is primed (Figure 2).

With the pump primed, thereservoir is at full pump dischargepressure. Flow through the recircula-tion port is minimal as pressures arenearly balanced, and in some designsthe port functions as an auxiliarycutwater – thus reversing flow direc-tion. Efficiencies on this type ofpump approach efficiencies on stan-dard centrifugal pumps for the samecapacity and head range.

Suction lifts up to 25ft are attain-able with this design, depending onimpeller diameter and speed. Closeclearances between the impellerdiameter and cutwater tip assurethat the liquid/air mixture is dis-charged out of the volute area andnot recirculated, which would cutdown on suction lift capability. Someself-priming pumps are designedwith replaceable or adjustable cut-water tips to compensate for wear.Thus, plant personnel can renewpriming lift ability without replacingthe entire pump casing.

Discharge Piping SystemsOne important fact about the

priming cycle is that the air evacuat-ed from the suction line is at a very

low pressure at the pump discharge.As noted earlier, little pressure ener-gy can be recovered as low densityair is passed through the impellerand separated from the liquid. A typ-ical plot of pressures observed at thepump discharge during the primingcycle is shown in Figure 3.

Traditional applications for self-priming pumps, such as dewateringan excavation or pumping out aflooded basement, normally don’trequire discharge piping (Figure 4a).Air from the suction line is easilyexpelled at the pump discharge untilthe unit is primed.

However, when the applicationrequires the addition of a dischargepiping system and incorporates acheck valve to prevent backflow orto stop water hammer in high verti-cal runs of pipe when the pump isshut down (Figure 4b), a problemarises during the priming cycle.When the pump is started, the checkvalve in the discharge line preventsthe air from being evacuated out ofthe suction line because it cannotdevelop enough pressure to over-come the head of liquid keeping thecheck valve closed. With no place forthe air to vent, the pump will notprime, which can lead to pump ormechanical seal damage.

Figures 4c and 4d illustrateoptions to prevent discharge piping

Figure 1. Internal construction of aself-priming centrifugal pump

Figure 2. The self-priming cycle

Figure 3. Typical pressures observedat pump discharge during primingcycle

Figure 4. Discharge piping systems

Figure 5. Example of discharge pipingsystem containing liquid loop or trap

206 The Pump Handbook Series

AIR

AIRLIQUID

MIXTURE

LIQUID

RECIRCULATIONPORT

AIR (A) (B) (C) (D)

AIRBLEED

AIR

UNSATISFACTORY

Vd

Vs

PRESSUREAT

PUMPDISCHARGE

FULL

PR

IME

LIQ

UID

REA

CHES

HO

RIZO

NTA

L RU

N A

T PU

MP

SUCT

ION

PUMPING TIME

AIR OUT LIQUID OUT

PRIMING PUMPING

Page 206: Centrifugal Pumps Handbook

system backflow yet allow the self-priming pump to evacuate air fromthe suction line. The air bleed line inFigure 4c should not be installedbelow the liquid level or contain anyliquid traps to impede air flow fromthe pump. The air release valveshown in Figure 4d allows the air toescape and seal once the pump isprimed.

Discharge piping systems con-taining liquid loops or traps similarto those shown in Figure 5 should beavoided. If the air occupying thepump and suction line volume (Vs)cannot be added to the air in the dis-charge volume (Vd) without exceed-ing the low pump discharge pressureduring the priming cycle, provisionsshould be made to allow the suctionair to vent as in Figures 4c and 4d.

Suction Piping SystemsAs applications for horizontal

self-priming pumps have expanded,suction piping systems have pro-gressed from a simple suction hosein a ditch to more complex pipingarrangements designed for transfer-ring liquids from tanks or rail cars toother storage facilities.

The suction system shown inFigure 6 is a piping arrangement fortank car unloading operation. Atstartup, full discharge pressure isexperienced as the pump is filledwith liquid. The unit continues topump the initial volume of suctionliquid into the discharge system, andwhen the air in the suction line isencountered, the discharge pressuredrops as the pump enters the prim-ing mode. The location of the initialvolume of pumped liquid in the dis-charge piping system should be

reviewed, as previously discussed, tobe sure it does not present a restric-tion to the evacuation of the suctionpiping air. Again, a bleed line or airrelease valve may be necessary.

VortexingVortexing is the introduction of

air into a pump due to insufficientdepth of liquid above the pump suc-tion line. High velocity flow patternsat the suction pipe entrance inducevortices or whirlpools in the liquid,and these open up a channel allow-ing air to enter the pump.

When a tank is pumped-down toits lowest level, vortexing may occur.At this point in the pump operation,the discharge pipe system will becompletely filled. To continue pump-down below this level, the air mustbe able to pass through the pump atreduced discharge pressures. As theliquid flow through the pump isreduced from the air drawn into thesuction, the vortexing will subsideand the pump will reprime itself pro-vided the air can pass through thepump. Upon re-prime, full flow willbe realized and the vortexing willreappear. This alternate primereprime cycle should be avoided sinceit can lead to premature bearing orseal wear if the frequency is toohigh.

This problem can exist in otherself-priming centrifugal pumps. Sub-mersible installations can experi-ence difficulty if air enters the pumpsuction from vortexing and cannotexit because of the closing of a dis-charge check valve due to reducedpressure.

Suction VolumeThe volume of air to be evacuat-

ed from a suction piping systemshould be examined to prevent pumpand/or mechanical seal problems.During the priming cycle heat isbeing added to the fluid in the pumpreservoir from the recirculation ofthe priming liquid. In most cases themechanical seal is being cooled bythe pumped liquid, and in the prim-ing cycle the heat being generated bythe seal faces is also being absorbedby the priming liquid. Duringextremely long priming cycles andhigh suction lifts the priming liquid

can evaporate. This means the pumpnever reaches prime, a situation thatcan result in mechanical seal failure.

Data on most commercial self-priming pumps include priming timecurves in addition to head-capacitycharacteristics. They show the timerequired to evacuate vertical suctionlines as a function of the distance ofthe pump above the liquid level.These characteristics are commonlyreferred to as lift curves.

Lift curve characteristics arebased on a liquid with a specificgravity of 1.0. If different specificgravity liquids are to be pumped, anequivalent lift should be used todetermine the priming time.

equivalent suction lift = suction liftx specific gravity

Increasing the suction pipediameter to reduce friction losses, orincluding long runs of horizontalpipe to be evacuated, lengthens thetime required for priming.

Increasing the vertical suctionpipe size results in an increase inpriming time proportional to thesquare of the pipe diameters. Hori-zontal runs of suction pipe and vary-ing pipe diameters require anumerical integration of the liftcurve based on the location of thediameter changes and horizontalruns to determine the resultant prim-ing time. If system priming timesexceed the maximum times shownon the lift curve, the manufacturershould be consulted.

In some instances with largesuction volumes, an auxiliary linecan be installed to add make-up liq-uid to the pump to prevent the prim-ing fluid from boiling off.

NPSHA/Suction LiftsOne important item to check on

the suction piping arrangement of aself-priming centrifugal pump is thenet positive suction head available (NPSHA). While NPSHA should bereviewed in all pumping applica-tions, it is especially necessary onpumps operating with a suction lift.NPSHA is the absolute pressureavailable to push the liquid throughthe suction piping up to the pumpsuction.Figure 6. Suction piping system for

tank car unloading

The Pump Handbook Series 207

Page 207: Centrifugal Pumps Handbook

NPSHA is defined as:

NPSHA = PBAR - PVP - PFR+ PHGT

Where PBAR = barometric pressurePVP = liquid vapor pressurePFR = suction line pressure

drop due to friction PHGT= height (pressure) of

liquid surface rela-tive to the pump

Note that all the terms makingup NPSHA are pressure terms andcan be expressed in feet of liquid. Inthe case of a suction lift, the last term(PHGT) would be negative as thepump is above the liquid surface.

Noting that PBAR (barometricpressure) at sea level is 34ft of water;with high suction lifts, the NPSHAdeclines rapidly even before thevapor pressure and friction lossterms are deducted.

The resultant NPSHA mustalways be greater (usually with amargin of 2-3 ft) than the NPSHR ofthe pump to prevent cavitation.NPSHR, the net positive suctionhead required, is a characteristic ofthe particular pump. It is determinedby test and shown on pump perfor-mance curves as a function of flowrate. NPSHR can be viewed as ameasure of the ease of moving liquidthrough the pump, and it increaseswith flow rate and speed.

Air LeakageWhile leakage is not normally

considered in piping design, it isbeing mentioned here because it is a

primary source of problems in self-priming pump performance. While aself-priming pump can produce afairly high vacuum, it is not designedto handle large volumes of air.

A small suction air leak at a highsuction lift will, in most cases, pre-vent the pump from ever reachingprime. A self-priming pump with a3” vertical suction line and a 10 ftsuction lift that can prime in 30 sec-onds has an average air handlingcapacity of only approximately 1cfm. The equivalent leakage area toflow 1 cfm at a 10 ft suction lift vacu-um is only .002 in2!

For this reason gaskets and sealsshould be in good condition to pre-vent air leakage during the primingcycle. Packing is not recommendedfor use in sealing self-priming pumpssince it is prone to leakage under thenegative operating pressures encoun-tered during priming. If priming dif-ficulties are experienced and airleakage is suspected, a commonpractice is to isolate the suction pip-ing system from the pump and checkthe pump for blank vacuum. Blankvacuum gauge readings should begreater than the suction lift require-ments. Typically, they run in the 20-25 in.Hg vacuum range.

SummaryHorizontal self-priming centrifu-

gal pumps have all of the inherentadvantages of standard centrifugalpumps such as:

• low initial cost• high capacity

• high suction lifts• relatively high discharge heads• ease of installation• ease of operation• ease of service • ability to handle dirty or solids

laden liquids

In addition, a self-priming pumpcan evacuate the air in the suctionline prior to pumping liquid. Howev-er, while it can produce high lift vac-uums, it cannot discharge theevacuated air against any back pres-sure.

The key to a successful installa-tion is to match the pump to theapplication, then match the pipingsystem to the pump. Avoid the twocommon pitfalls of self-primer instal-lations – namely regarding the pumpas an air compressor and imposingback pressure on it during the prim-ing cycle, and allowing air leakage inthe suction line or pump seal area.Following these guidelines will resultin relatively trouble free operationand allow the pump to systematical-ly remove the air out from the suc-tion line. This will result in efficientpumping of liquid, and upon shut-down, the pump will retain enoughliquid to repeat the cycle.■

Ray Petersen is Manager of Engi-neering for the Fybroc Pump Divisionof Met-Pro Corporation (Telford, PA).He holds a master’s degree in mechani-cal engineering from Drexel Universityand has more than 25 years of experi-ence in the centrifugal pump industry.

208 The Pump Handbook Series

Page 208: Centrifugal Pumps Handbook

CENTRIFUGAL PUMPS

HANDBOOK

The Pump Handbook Series 209

Know the Inside Story ofYour Mag Drive Pumps

Accurate monitoring of temperature and pressure is key to reliable operation.

By Harry Schommer

Editors Note:Mr. Schommer, of Waldkraiburg,

Germany prepared this article forPumps and Systems. To assist our U.S.readers not entirely familiar with SI(metric) units, we have included an eng-lish/metric conversion chart at the endof this article.

T he pump shafts in seallessmagnetic driven pumps areheld in place by sleeve bear-ings. Therefore, these bear-

ings must be located in the pumpedliquid. Today, the most commonbearing material is pure silicon car-bide. Silicon carbide bearings workwithout problems in low viscosityliquids such as chemicals, hydrocar-bons, solvents, acids, all kind ofhydroxides and also in abrasivepumpage. With an additional dia-mond-layer, the material providesdry running capabilities.

The widely used term “processlubricated bearings” is not quite cor-rect since there are no lubrication

grooves applied andthere is no defined flowthrough these bearingsprovided in conven-tional sleeve bearingdesign. Liquids such as propane, ethyleneoxide or methylenechloride provide nolubrication capability.Similar to the situationbetween the faces ofmechanical seals, onlya stable fluid film isrequired between theslide faces. The stabili-ty of this film dependson temperature andpressure. If tempera-ture rises inside thepump above vapor tem-perature of the pumpedliquid, vaporizationwill break down this film. Underthese conditions, the bearings willrun dry and fail sooner or later. Areliable temperature monitoring sys-

tem is required to avoidthis situation.

Besides vaporizationof fluid inside the pump,dry running of an emptypump is the worst operat-ing condition. Because ofthe starved suction, thereis no flow to any part ofthe pump. Although thediamond-layer of the SiC-bearings will tolerate thissituation because no hy-draulic loads are acting,the built up heat whichcannot be dissipatedbecause of the starved

internal circulation will lead todemagnetization of the coupling.

Temperature Rise in SingleStage Volute Casing Pumpswith Magnetic Couplings

Internal CirculationSealless pumps with magnetic

couplings and metallic containmentshells generate eddy currents thatlead to heat and cause temperaturerise of the pumped liquid in the con-tainment shell. In order to preventthis, heat must be dissipated throughan internal cooling flow. This coolingflow – branched off as a partial flowfrom the main flow and led throughthe gap between internal rotor andcontainment shell – is shown inFigure 1.Magnetic drive centrifugal pumps with metallic

containment shells

40

35

30

25

20

15

1010 20 30 40 50

TE = TE’=NPSH - stable

Minimum flow Qmin Q [m3/h]

TA

p3p4

T’E

TE

PS

PA’ TA

Tcs’ Pcs

Tcs

∆T

Rotor back vanes

t [°C]

Abb. 1Figure 1. Internal circulation, temperature behavior

Page 209: Centrifugal Pumps Handbook

The circulation flow is drawnfrom the discharge side behind theimpeller, led into the chamberbetween the slide bearings andthrough the pump shaft via the rotorback vanes, and returned to the dis-charge side. This arrangement pres-surizes the slide bearings and thecontainment shell with nearly thefull discharge pressure, and helps toavoid flashing of the liquid in thisarea caused by heating up the prod-uct. Where the temperature increaseis critical, the maximum pressure P4

prevails. It should be noted that thereis no heated liquid flowing back tothe suction side or impeller. There-fore, no negative influence on theNPSH required will occur. For thistype of pump, handling of volatileliquids is not a problem.

In pumps without rotor backvanes or rear impellers, the internalcooling flow is driven by the pressuregradient within the pump from dis-charge side to suction side, i.e. backto the impeller eye. In this case, prob-lems may arise in pumps handlingvolatile liquids. Sufficient NPSHreserves must be available to accom-modate the heat-conditioned rise ofthe pump’s NPSH value. Exact tem-perature measurements of the cool-ing flow after passing the magnetarea are also impossible.

Temperature Rise, MinimumFlow Conditions

Figure 1 further displays thetemperature behavior in a volute cas-ing pump with a magnetic coupling.The pump size is 50/200, 2900 rpm;magnetic losses are 3,0 kW and thepumped liquid is water. Whenreviewing the temperature curve itmust be considered that the flowleading through the magnet chamberis dependent only on the geometry ofthe rotor back vanes and on thespeed. This means that, independentfrom capacity and differential head,a stable circulation flow exists thatadopts the dissipated heat and leadsit into the main flow. Since the mag-netic losses of a given magnet cou-pling will not change duringoperation at constant speed, analmost stable temperature increase∆T is produced in the range to theright of the minimum flow. However,

if the minimum flow drops below,temperatures will rise remarkably.This is the reason why these pumpscannot be operated against closed discharge valve. Experience hasshown that most of the slide bearingdamages are a result of neglectingthis fact. If process conditions dictatethis operation, a bypass line must beinstalled from discharge to suctionvessel.

Numerous temperature mea-surements on magnetic drive pumpswith different sizes, rotor diameters,containment shell materials andspeed have proved a direct relationbetween the capacity Q, the magnet-ic losses Pv and the temperature ∆Tcs

inside the containment shell. Theserelationships are displayed in Figure2, based on water at 20ºC. Determi-nation of the actual temperatureincrease inside the containment shellfor other liquids depends on theproduct.

∆Tcs,product = ∆TH20 .

spec.heatH20 . densityH20 [ºC]spec.heat product density product

Knowing the inlet temperatureTE, the containment shell tempera-ture Tcs, product is determined as fol-lows:

Tcs, product = TE + ∆Tcs,product [ºC]

Maximum AllowableContainment ShellTemperature

When handling volatile liquidsor products with a vapor pressurethat complies with the pump suctionpressure Ps, the relation betweencontainment shell temperature, con-tainment shell pressure and boilingpoint of the liquid must be taken intoconsideration. Only when the operat-ing conditions are not beyond theboiling point – that is, liquid is notflashing inside the containment shell– can safe operation be guaranteed(Figure 3). The condition point mustalways be in the liquid state.

Basically, the pressure rise ∆Pcs

inside the containment shell duringoperation must always be higherthan the heat-conditioned rise ofvapor pressure ∆PD of the product.To determine the maximum allow-able containment shell temperatureTzul, the vapor pressure curve of theproduct must be available (Figure 3).The boiling temperature TD can betaken from the intersection pointbetween containment shell pressureat duty point of the pump and thevapor pressure curve. By adding acertain safety margin ∆Ts, the maxi-mum allowable containment shelltemperature Tzul can be determined.However, it must always be higherthan the calculated containmentshell temperature Tcs in order toavoid vaporization. This means ahigh pressure rise ∆Pcs in the con-tainment shell area provides a highersafety factor.

Pumps with internal circulationfrom discharge to discharge side,through back vanes or rear impellers

210 The Pump Handbook Series

60

58

56

54

52

50

48

46

44

42

40

38

36

34

16

15

14

13

12

11

10

9

8

7

6

5

4

3

2

1

5 10 15 20 30 40 50 70 90 120 150 200 300 U.S.G.P.M.

1 1.5 2.5 4 6 8 10 15 20 30 50 70 m3/h

[°F] [°C] ▲▲Tcs, H2O

parameters Pv [kW]

▲▲T-values based on H2OO recommended Qmin H2O

17kW

8.3kW

4.3kW

2.7kW

2.0kW

5.6kW

Thermal stable(▲▲T const.)

Figure 2. Temperature increase

∆Pcs

Pcs

PSTE

∆TS

∆TProduct

∆PD

TD

Liquid

p[bar]

Vapor

Vapor pressure PD

T[°C]Tcs Tzul.

Figure 3. Containment shell, vaporpressure curve

Page 210: Centrifugal Pumps Handbook

(Figure 1), are pressurized in the con-tainment shell area by approximately80% of the differential head plus theinlet pressure Ps. The containmentshell pressure for this pump designcan be calculated as follows:

Pcs = Ps + ∆Pcs [bar]The pressure increase ∆Pcs

depends on the rated differentialhead H and the density of the liquid:

H . p . 0,8 [bar]∆Pcs = 10,2

H[mLC], ρ[kg/dm3]

Pump series with circulationfrom discharge side to suction sidehave lower ∆Pcs values. Exact valuescan be learned from the pump manu-facturer or taken from the pump datasheet.

Practical ExampleA standard chemical pump with

metallic containment shell (HastelloyC) and partial circulation as indicatedin Figure 1 is required for handlingammonia (NH3). Service conditionsare as follows:

• Liquid: NH3, TE = 0 ºC, Ps = 4,4 bar, Q = 0,66 kg/dm3, C = 0,492 g/cal/ºC

• Differential head 120 mLC• Capacity 40 m3/h• Coupling losses Pv = 2,7 kW

The expected containment shellpressure Pcs is to be determined first:

Pcs = Ps + ∆Pcs

Pcs = 4,4+120 . 0,66 . 0,8 = 10,6 bar10,2

Based on the vapor pressurecurve for NH3 (Figure 4) and on thecalculated containment shell pres-sure of 10,6 bar, a boiling tempera-ture of +25ºC is given which mustnot be exceeded during operation.For final determination of allowabletemperature Tzul, the actual expectedcontainment shell temperature at thethermal stable minimum flow of 8m3/h (Figure 2) must be calculated:

Tcs = TE + ∆Tproduct

= 0+4 . 1 . 1 = 12,3ºC0,492 0,66

Considering the boiling temper-ature of 25 ºC and the actual temp-erature of 12,3 ºC, the allowablecontainment shell temperature Tzul

can be defined at 20 ºC.

Temperature Monitoring

GeneralCentrifugal pumps with magnet

coupling are not considered electri-cal equipment. Level detectors ortemperature controldevices are not requiredby government authori-ties for these pumps,even when they areinstalled with explosionproof motors in haz-ardous areas. However,application experienceshave shown that themain reason for opera-tional troubles – besidesworn antifriction bear-ings – is the practice ofexceeding the allowablecontainment shell tem-perature. Consequences of excessive

temperature rise arecavitation in the area ofthe containment shelland dry running of theslide bearings due toproduct flashing. Caus-es for this can be opera-tion below minimumflow or against a closeddischarge valve, block-ing of the inner rotor,clogged circulationholes and/or solid parti-cles between rotor andstationary shroud.

Therefore, it is

essential to monitor magnetic drivenpumps with temperature probes toensure an automatic switch offbefore serious damage occurs.

Temperature ProbesFor permanent control, resis-

tance temperature probes are thepreferred method of monitoring con-tainment shell surface temperature,although this kind of protection isnot activated under dry running con-ditions. Nor can containment shellrupture by the outer magnets result-ing from worn antifriction bearingsbe avoided.

Typically, the probes work witha measurement rheostat of platinum,showing an electrical resistance of100Ohm at 0ºC. Temperaturechanges at the measuring point leadto a change of the resistance and, inturn, to a change of the voltage. If thepreadjusted temperature limit isexceeded, the voltage change switch-es off the driver through a connectedcontroller.

Figure 5 shows a standard probeadaptable for containment shells. Ithas a flat bottom for sufficient con-tact to the containment shell surface.The element is located directly onthe bottom of the probe. Continuouscontact between the probe and con-tainment shell surface is guaranteedby an integrated compression spring.

Another type of temperatureprobe, located in the internal recircu-lation flow, measures the fluid tem-perature after passing the magnetarea. This system works sufficientlyif the pump is properly filled withliquid. It protects against exceedingthe boiling point of the liquid in the

The Pump Handbook Series 211

30

20

108

6

4

2

10,6

4,4

-30 -20 -10 0 +10 +20 +30 +40 +50+25

Vapor pressure NH3

PD [bar]

Figure 4. Vapor pressure curve, NH 3

Figure 5. Temperature probe (PT 100)

connection headto motor circuit

Spring loaded

Ceramic sandCementPT 100-element

Bottom

internal circulation

adapter pieceSpring

Extension

sealing

Protection tubeShroud protection

drive magnet

shrouddriven magnet

Page 211: Centrifugal Pumps Handbook

area of the containment shell causedby excessive temperature rise.

The problems of protectingpumps against dry running throughtemperature probes have alreadybeen pointed out. Given an emptypump, i.e., under dry running condi-tions, it has been proven that the con-tainment shell surface temperatureT1 in the center of the magnet cou-pling (Figure 6) deviates remarkablyfrom the surface temperature on themeasuring point T2. The reason forthis are the eddy currents occurringin the center of themagnets that rapidlyincrease the contain-ment shell tempera-ture. The temperatureprobe at measuringpoint T2 cannot detectthis temperature risein time because of thebad thermal conduc-tivity of the shellmaterial (18.10.CrNior Hastelloy). Conse-quently, these type ofprobes are not able toprotect magnetsagainst overheating ofan empty pump.

To obtain reliablemeasurements withthis type of monitor-ing, the pump must bevented and the probelocated at the reverseinternal cooling flow(after passing the mag-

net area). This is only possi-ble in designs with rotorback vanes or rear impellers.If the temperature probe islocated in the precirculationflow (this is the case whencirculating to the suctionside), serious problems occurwhen pumping volatile liq-uids. With this setup, anincrease in temperature isindicated only when thecomplete pump has alreadybecome hot – too late to pre-vent flashing in the magnetend.

MAG-SAFETemperatureMonitoring

Figure 7 shows therecently developed MAG-SAFE tem-perature monitoring system that isable to read the temperature directlyon the heat source. It records theactual temperatures occurringbetween the magnets inside the con-tainment shell and converts theminto a linear output of 4 to 20 mA.Therefore, it is possible to preadjustthrough a trip amplifier any shut-offtemperature within the range of -5 to250ºC.

Compared with common tem-

perature control (Figure 5), the MAG-SAFE system features:

• Extremely fast reaction time toall temperature rises (Figure 6) , andswitches off under dry running con-ditions.

• Since the containment shell is aheat source because of the eddy cur-rents induced in it, a temperaturerise can be detected before the liquidtemperature in the containment shellis remarkably affected. Exceedingthe boiling point can be prevented ifthe limit temperature is correctly set.

• Excessive temperatures at con-tainment shell surface within the Ex-area can be prevented.

• The direction of the internalcirculation flow has no influence onthe temperature indication.

• Worn out ball bearings causeeccentric run of outer magnets andwill lead, if not detected, to erosionat the protection device and contain-ment shell rupture. With the MAG-SAFE system, drive magnets cut theconnection wire 3 if such a conditionis not recorded in time, and thepump is switched off before seriousdamage can occur.■

Harry J. Schommer is Chief Engi-neer of Dickow Pumpen KG in Wald-kraiburg, Germany, with responsibilityfor the design and development of thecompany’s sealless magnetic drivepumps. He has authored several paperson pumps and shaft-sealing systems,and conducted seminars for pump usersin Germany, Europe and the Far East.

212 The Pump Handbook Series

500

450

400

350

300

250

200

150

100

50

2060 120 180 240 [sec]

T1 , T2 [°C ]

Coupling design:PNK : 21,5 kW / 2900min-1

PV : 2,0 kW

T1 = MAG SAFET2 = PT 100

T2

Figure 6. Containment shell surface temperature

Figure 7. MAG-SAFE temperature monitoring

10

12

43

117

8

6

9

512

SI Unit Approximate Resulting Conversion Factor English Unit

kW 1.34 hpm3/h 4.4 US gpmºC 9/5 ºC + 32 ºFmLC 3.28 feet(of head)(meters liquid column)

1. Containment shell 7. Connection piece 2. Outer magnets 8. Transmitter3. Thermocouple wire 9. Connection clamps 4. Thermocouple 10. Cable inlet5. Connecting block 11. Protection device6. Thread connection 12. Bearing bracket

T1

Page 212: Centrifugal Pumps Handbook

Pump Rebuilding at Avon

This refinery’s successful program meets strict emission standards and enhances pump reliability at the same time.

By Stephen C. Rossi and John B. Cary

Most states have adoptedregulations that limit fugi-tive hydrocarbon emis-sions from mechanical

seals in centrifugal pumps. In Cali-fornia, limits as low as 100 ppm havebeen imposed. Users, faced with fewchoices to meet these strict stan-dards, have turned to dual seals andsealess pumps to comply. Manyusers, however, have found that theycan meet current fugitive emissionlimits with single seals by payingcareful attention to detailed retrofitand repair procedures. This articlediscusses the rigors required torebuild, maintain and operate pumpssuccessfully – in most cases, with asingle seal. These techniques alsoenhance pump reliability and havebeen applied to more than 100pumps in harsh refinery environ-ments.

Background

Pump History at Avon

The Avon Refinery is more than80 years old. In 1913 crude oil pipestills were built by a group of SanJoaquin Valley oil producers on theCarquinez Straits near Martinez, Cal-ifornia. Shortly after building awharf for receiving the crude oil,they commenced construction ofwhat is now the Avon Refinery a fewmiles away.

The oldest operating processplants date back to the late 1930s.Several facilities were built at theonset of World War II to support thedefense effort. They were equippedwith high speed centrifugal pumpsalmost exclusively. New plants havebeen added every decade since, but

the average age of a centrifugal pumpfor the whole complex is almost 20years.

Pumps of almost every typewere installed over the years. Manywere converted from packing tomechanical seals without regard toshaft flexibility and operating point.Over time poorly assembled andmaintained auxiliary piping andflush systems were added, modifiedor abandoned. In addition, unclear ornonexistent operating procedures leftoperators to use their judgment. Allof these factors contributed to poorpump/seal reliability.

Evolution of BAAQMD Regulations for Pumps

In the early 1970s the CaliforniaAir Resources Board (CARB) mandat-ed emissions standards for refineries.Two regional agencies were formedto monitor enforcement of the newstandards – the Bay Area Air QualityManagement District (BAAQMD),and the South Coast Air QualityManagement District (SCAQMD).These agencies were also authorizedto develop compliance standards fortheir jurisdictions. Historically, these

standards have been much morestringent than federal emissions stan-dards.

The first hydrocarbon emissionsstandards for pumps limited releaseof Volatile Organic Compounds(VOCs) to 10,000 ppm (parts per mil-lion), expressed as methane. TheBAAQMD also required no less fre-quent than quarterly monitoring ofpump seals. The district also mandat-ed detailed record keeping. Fifteendays were allowed to repair pumpsfound over the emissions limit. Theequipment could not be returned toservice until the emissions limit wasmet. These limits were attained inmost cases by replacing mechanicalpacking with single pusher-typeshaft mounted mechanical seals.

As pumps were brought intocompliance, the regulators began to“ratchet down” (their term!) on refin-ers. The regulation became increas-ingly stringent and, beginning in1993, requirements went into effectregulating the percentage of equip-ment that could continue to operateand be put on a future or turnaroundrepair list (Table 1).

1992 & prior 10,000 ppm 15 days no limit

Jan. 1, 1993 1,000 ppm minimization 10% Spared equip. in 24 hours, may not be on repair in 7 days T/A list

July 1, 1993 same as above same as above same as above Nat Gas added

Jan. 1, 1995 same as above same as above same as above Methane

Jan. 1, 1997 500 ppm same as above 1% Nat Gas added

TIME TO %WAITINGYEAR LEAK STD REPAIR T/A REPAIR REMARKS

Table 1. Summary of BAAQMD regulations for pumps

CENTRIFUGAL PUMPS

HANDBOOK

The Pump Handbook Series 213

Page 213: Centrifugal Pumps Handbook

As these emissions standardswent into effect, it became more dif-ficult – and in some cases impossible– to maintain emissions levels in old-er style pumps. Slender shafts andlong spans between bearings createdtoo much shaft deflection at the sealfaces. Small seal chambers and inad-equate clearances added to the diffi-culty of retrofitting pumps withmodern mechanical seals. At thesame time, seal technologies wereevolving. Cartridge mounted sealsand requirements for larger sealchambers drove Avon managers toembark on a major fugitive emis-sions reduction program.

Initial ApproachIn January 1993 the BAAQMD

mandated that pump hydrocar-bon emissions levels be reduced to1000ppm. To meet this more rigor-ous standard, the refinery reliabilitygroup took the lead in identifying thescope of the program, with the intentof turning over the project to engi-neering for execution. The followingapproach was used for the initialscoping project:

• data collection and analysis• seal performance testing and

pump evaluation• vendor selection

Data Collection and Analysis

Data was collected and analyzedto develop a list of pumps requiringwork. At this point, the work itselfwas undefined. Formal fugitive emis-sions testing was done every quarter.Three years of data on approximate-ly 500 pump seals was analyzed toidentify all pumps that had failedtheir emissions test more than onceper year over that period. This infor-mation was compared with the refin-ery’s “bad actors” list and prioritizedaccordingly. Estimates and scheduleswere developed for the annual bud-getary cycle. Other data sources werereviewed to gain a thorough under-standing of the scope of the project.These included:

• pump importance and its effect on the process

• failure history, including mean time to repair

• maintenance cost history (total

and per repair)• bad actors lists• hydraulic performance (NPSH

and the difference between BEP and actual operating point)

• future hydraulic requirements

External information was alsocollected. A benchmarking studywas undertaken in other West Coastrefineries to determine what theirexperience had been with variousseal designs and seal vendors.

Seal Performance Testing and PumpEvaluation

The initial data indicated thatthere were about 150 pumps requir-ing some degree of retrofitting. Themajority of these pumps handled lowviscosity products such as propane,butanes and light gas oils. Based onthe scope of the project and the timeallotted, the decision was made torely on one seal manufacturer. Thiswould save the time involved in com-petitively bidding each seal, and itwould reduce the overall costthrough volume purchases.

Three identical pumps in thesame service were chosen for initialseal testing. These pumps wereselected because they were in lighthydrocarbon service and moving liq-uid near its vapor pressure. In addi-tion, the pumps were in a unit thatwas going into a maintenance turn-around, providing an opportunity torebuild all three pumps to identicalspecifications.

Three pump vendors were invit-ed to participate in the project. Theywere asked to submit proposals toprovide their “best available controltechnology” (BACT), and they weregiven the specifications for the testpumps. The objectives for the testproject were to:

• Establish a control technology for light hydrocarbon services to be used on pumps that must be retrofitted to meet future emissions requirements.

• Provide single seals designed to meet stringent emissions standards, avoiding expensive and complicated dual seal installations.

• Document this technology and

performance so it could be applied to retrofits as well as new installations.

A technical support agreementfor the test program was developedfor the participating seal vendors,stipulating the obligations of each toprovide technical support and assis-tance as a full partner in seal selec-tion, retrofits and testing.

Each vendor assigned a field ser-vice engineer to participate in theoverhaul, installation, start-up andfield monitoring of the test seal. Thetest pumps were meticulously over-hauled and carefully installed. Theseals were tested over a three monthperiod, with the following datarecorded daily:

• emissions readings• suction and discharge pressure• pumped fluid temperature• vibration (radial and axial)• seal flush pressure• quench steam flow

Vendor Selection

At the end of the seal trial peri-od, field test data were comparedwith results of the alliance selectionprocess led by the plant’s purchas-ing group. Vendor performance wasweighed against several pre-estab-lished performance criteria, such asdegree of technical support, seal per-formance, response time, level oftechnical expertise and experience.Long term alliance agreements werethen drafted, budgets prepared andproposed schedules developed.

Project Approach

Defining the Project

After establishing the breadth ofthe project, Avon officials initiated aformal program to implement thework. The program was divided intophases corresponding to annual bud-get cycles. A project core team wasassembled, consisting of a projectmanager, project engineer, full timevendor technical representative,draftsperson, clerk, mechanical con-tractors, pump alliance partner and abuyer from the Purchasing Depart-ment. Various area maintenance andoperating personnel became ad-hocparticipants, depending on the loca-tion of the retrofit work.

214 The Pump Handbook Series

Page 214: Centrifugal Pumps Handbook

The charter of the project teamwas as follows:

• Develop reliable technology for pump emissions control to comply with the 1993 through 1997 emissions limits.

• Implement this technology as proactively as possible to avoid penalties and fines for non-compliance.

• Complete detailed designs required to achieve reliable 1993-1997 emissions com-pliance.

• Provide project management support for all phases of equipment upgrades.

• Develop and evaluate equip-ment hydrocarbon emissions data to prioritize and schedule optimal cost effective repairs and upgrades.

• Integrate the activities of the project with that of the main-tenance and operating depart-ments to minimize disruption in the operation of the refin-ery.

• Recommend and coordinate equipment upgrades meant to improve equipment reliability in conjunction with emissions compliance modifications.

Candidate Pump Evaluation

Emission levels were measuredon all pumps identified as havingpotential VOC compliance problems.Qualitative data on all VOC programpumps were surveyed to assess theneed for upgrades to the sealing sys-tem to meet 1997 regulations.

Pumps found to be above the500 ppm limit for 1997 were consid-ered for possible upgrade. The com-pany’s third party contractor foremissions compliance used a database system to generate queries onall pumps with emissions levels at orabove the limit. Detailed analysis ofeach pump system produced a break-down of anticipated upgrades andreplacements (Table 2).

Proactive Approach to Problem Pumps

A proactive approach was takenbecause an in-kind, reactive repairprogram would have increased costto the company and created majorrepair backlogs when more stringent

emissions limits went into effect.The cost of emissions related

pump repairs made in 1992, 1993and 1994 is shown in Figure 1. In1994, 40 pumps required emissionsrelated repairs. This was down sub-stantially from 92 in 1992 and 64 in1993. The total cost for all repairsduring this three year period wasmore than $1 million.

The following projections illus-trate the potential for savings with aproactive vs. reactive approach toemissions compliance. (Repairs wereexpected to remain constant through1996, then increase by 50% in 1997when the emission limit was reducedto 500 ppm).

Savings over this three year period was estimated to be at least$175,000 as shown in Figure 2. Acomplimentary maintenance costsavings was also expected due to

increased reliability.In conjunction with emissions

history, mean time between repair(MTBR) and hydraulic performancerequirements were evaluated for pos-sible upgrade.

Experience had shown that in-kind repairs of equipment that failedemission monitoring would, in mostcases, not comply with 1997 levelsdue to pre-existing problems such aspipe strain, unstable foundations andpoor suction conditions. These condi-tions were corrected as part of theupgrade process. This approachreduced chronic mechanical seal fail-ures and subsequent emissions,improved long term reliability andlowered total life cycle cost of theequipment.

Tremendous effort was expend-ed in identifying pumps to be includ-ed in the program. Data from anumber of sources was scrutinized tojustify each pump’s inclusion. Thiswas an important exercise because itprioritized the upgrade sequence(shortest MTBR to longest MTBR),identified detailed scopes of work,and scheduled each retrofit to coin-

The Pump Handbook Series 215

Seal Upgrade only 96Power End Retrofit 45Complete Replacement 5No Change 24Total 170

Number Recommendation of Pumps

Table 2. Recommended modifi cations

$200,000$180,000$160,000$140,000$120,000$100,000$80,000$60,000$40,000$20,000

$0

Figure 2. Project pump repair costs

1995 1996 19971995: 25 emissions repairs per year at $5000 each

$125,0001996: 25 emissions repairs per

year at $5000 each$125,000

1997: 40 emissions repairs per year at $5000 each

$200,000

TOTAL COST $450,000

“Reactive” Approach (in-kind repairs on as needed basis)

1995: 25 emissions repairs per year at $5000 each

$125,0001996: 15 emissions repairs per

year at $5000 each$ 75,000

1997: 15 emissions repairs per year at $5000 each

$ 75,000

TOTAL COST $275,000

“Proactive” Approach(continuation of modifications

and thermal oxidizer installations)

Figure 1. Annual cost of emission

$500,000

$400,000

$300,000

$200,000

$100,000

$01992 1993 1994

Reactive Program CostsProactive Program Costs

Page 215: Centrifugal Pumps Handbook

cide with other refinery activities.At the end of this process, a well

defined list of pumps targeted forupgrade emerged. The plan was com-municated to all affected refinerypersonnel and was used as the basisdocument for all project work.

Preparation of Specifications

Specifications for Pump and Seal Purchase

The project team immediatelyset out to establish minimum re-quirements for pump and seal speci-fications. A series of meetings washeld with both the seal and pumpalliance partners, as well as withcompany machinery specialists, tocreate project specific specifications.The basis for the specifications wasAPI 610 with clarifications in theareas of fits and tolerances. A draftversion of API 682, (CentrifugalPump Shaft Sealing Systems forRefinery Services; first draft 9/92)wasalso used for guidance. The intentionwas to create an environment for theseal that would allow it to function asintended. These specifications werelater adopted company-wide.

Examples of Critical Fits

Areas of concern and additionalrequirements to API 610 are tabledbelow:

Additional requirements:

• seal chamber register (radial) to shaft within 0.001 T.I.R.

• component match impeller to shaft fit to achieve a goal of 0.000 inches tight to 0.001 inches loose

• component match of rolling element thrust bearings to achieve 0.004 inches maxi-mum axial float

• shaft sleeve bores equal to the maximum diameter of the shaft with a tolerance of +0.0010 inches to -0.0000 inches

Modifications:

• Squareness of seal chamber face register to shaft axis was reduced from 0.002 to 0.001 T.I.R.

• Sleeves will have a relief cen-tered axially and the minimum sleeve thickness can be 0.090

inches within the relieved area.

It was agreed that these specifi-cations would be the minimum qual-ity level expected.

Repair and Installation Specification

The quality of pump installa-tions – including foundation prepara-tion, grout (or lack of it), pipingstrain, alignment and other key fac-tors – varied considerably through-out the plant. No company repairand installation standard existed oth-er than the original equipment man-ufacturer (OEM) guidelines and afew rules of thumb. Consequently,the project team created a repair andinstallation specification. This docu-ment covered setting of new pumpsand drivers, rebuilding of pumps,installation of retrofit kits, and sealflush, vent and drain connections.

The following are importantfocus areas addressed in the specifi-cation:

• “As found” pipe strain effectswere checked and recorded (Figure3) prior to pump removal, using alaser alignment tool attached to thecoupling. These were checked againwhen final piping connections weremade.

• Existing pump base was checked for voids and flatness (Figure 4). Pressure injection grouting and field machining of mounting pads were carried out when tolerances were exceeded.

• On non-retrofit pumps, new 17-4PH pump shafts were fab-ricated and all fits reclaimed to tolerances in the new pump specification.

• All pumps, including retrofits, were fitted with close clear-ance carbon throat bushings to maintain seal chamber pres-sure.

• New dynamically balanced, multiple disk spacer couplings with register fits for the hubs and center section were pro-vided on all upgrades.

• Completed pump and seal assemblies were leak tested before field installation. This minimized the need for rework after the pump system was filled with process liquid.

Support Systems Specification andSelection

Existing piping and seal systems

Figure 3. “As found” pipe strainrecord form

DATE ___________________MACHINIST_______________APPROVED BY ____________EQUIPMENT ID ___________UNIT NUMBER ____________

DIS-ASSEMBLY

ANGULARHOT ALIGNMENT

0°90°180°270°360°

ANGULAR PARALLEL0°90°180°270°360°

COLD ALIGNMENTPARALLEL

ANGULARPIPE ALIGNMENT

L-RF-B

NO DSPLACEMENTTORSIONAL DISPLACEMENT

1/4 HOLE

1/2 HOLE

3/4 HOLE

PARALLEL

Figure 4. “As found” baseplate condi -tion

USE THE ABOVE TO MARK ANY AREASOF CORROSION, VOIDS, OR PITS ON THEBASE PLATE INCLUDING GROUT.USE THE FOLLOWING

V = VOIDC = CORROSIONP = PITS

LASERZERO POSITION

MOTOR PUMP

216 The Pump Handbook Series

Page 216: Centrifugal Pumps Handbook

were upgraded. Most of these pumpshad screwed connections (potentialemissions sources) throughout. Aspart of the upgrades for piping relia-bility, the following were addressed:

• All screwed connections on thepump casing or process piping werereplaced with Schedule 160 nipples;the nipples did not exceed 4” inlength and were gusseted in twoplanes, seal welded and flanged –eliminating screwed connections ifpossible (Photo 1).

• All tubing connections to theprimary seal were 1/2” 316 stainlesssteel with a wall thickness of 0.065”.Smooth radius bent 3/4 inch tubingwas used for thermosyphon coolingwhen required.

To reduce seal flange distortionand make installation and removal ofthe seal easier, the staff made surethat the final tubing connectionswere not more than 18 inches long.

Vent systems were provided onseal chambers in addition to thepump case to ensure filling of theseal chamber prior to startup. Theseventing systems included instruc-tions in the form of a venting proce-

dure tag installed on nearby piping.The message addressed both seal andpump venting (Photo 2).

Instrumentation

Only essential – or remote andunattended – pump seal systemswere instrumented. Typically, a levelswitch on a seal reservoir was theonly signal back to a control room.Most installations did not warrantremote readout instrumentationsince the operators were betterinformed of the general health of thepumps by observing them in person.

Secondary Containment (VRS)

A number of pump seal systemsused existing vapor recovery systems(VRS) for 100 percent containmentwhen it was available close to thepump installation. In many remotelocations total containment wasrequired, but VRS was not available.

Alternative Technologies

Where total containment wasrequired and the only choice was anitrogen pressurized dual sealarrangement, an alternative involv-ing new emissions control technolo-gy was used. This technology utilizedcompressed air passing through a jetejector to pull emissions from pumpseals or barrier fluid reservoir ventsthrough a flameless reactor that ther-mally oxidized the VOC’s to watervapor and carbon dioxide, as shownin Photo 3. Four units have beeninstalled with good success; the sys-

tem is 99.99% effective in reductionof VOC’s.

Each unit currently controlsemissions from 10 pumps. Futureexpansion up to 20 pumps is possi-ble. These systems are expected toavert as many as 15 emissions-relat-ed repairs per year. Many non-com-pliant pumps are scheduled to beconnected to a thermal oxidizer with-in the next year, possibly eliminatingthe need for further modifications.

Part 2Seal performance is affected by

numerous internal and externalforces. How a pump is sized for anapplication and how it is actuallyoperated have a significant impact onseal life. In fact, both of these factorscan shorten seal life in a pumpingsystem. Mechanical problems suchas misalignment, unbalance and flat-ness, as well as poor concentricityand perpendicularity are fairly wellunderstood and relatively easy tocontrol. Hydraulic forces, on the oth-er hand, are generally not as wellunderstood or recognized by person-nel responsible for daily pump oper-ation and maintenance. Further-more, it is usually more difficult toremedy a hydraulic problem since itmay often relate to the originaldesign of the system.

Shaft deflection and vibrationcaused by unbalanced hydraulicforces can be very destructive to apump and severely diminish seal life.Before embarking on a project toimprove seal performance, it is im-perative that the pump’s hydraulicperformance be verified. The closer apump operates to its best efficiencypoint (BEP), the longer the seal willlast. This has been demonstratedmany times in the field and wasrecently proven analytically in acomputer model specifically de-signed to predict seal life based on apump’s proximity to BEP. A discus-sion of the process for evaluatingpump hydraulics is included in theappendix to this article.

Developing AlliancePartnerships

To be lasting, an alliance rela-tionship must be profitable or benefi-cial for all parties. Partners need toPhoto 3. Thermal oxidizer system

Photo 2. Seal and pump venting procedure

Photo 1. Piping reinforcement detail

The Pump Handbook Series 217

Page 217: Centrifugal Pumps Handbook

take mutual responsibility to ensurethat the desired goals are achieved.One of the first steps the newlyformed Avon project alliances under-took was to develop well definedobjectives along with a mission state-ment.

Alliance Team Mission Statement

Adopt a fundamental philosophyof decreasing mechanical seal lifecycle costs through increased equip-ment reliability.

• Maximize equipment availability• Manage and document change ac-curately and completely• Improve data quality (new andexisting)• Obtain accurate process informa-tion from the owner/user• Analyze the root causes of failure• Maintain honest communicationabout failure by owner• Strive for total buy-in by manage-ment and staff – down to the last per-son

The alliance teams also devel-oped matrices to assess the benefit ofthe arrangement and continueimproving it:

Dollars 1) cost of new seal purchases 2) cost of seal repairs 3) inventory reduction 4) market share

Reliability1) number of repairs2) mean time between failures

Contractor Performance1) plant-wide pump survey status2) on-time delivery 3) failure analysis submittal

Company Performance1) appropriate paperwork 2) on-time payment 3) provides pump access

The key is to involve alliancepartners in all facets of project activ-ity.

Candidate Pump/Seal EvaluationBoth the pump and seal alliance

partners participated in all facets ofthe pump/seal evaluation.

• When a pump was added to thelist, based on emissions survey, thepump vendor and company repre-sentatives interviewed operationspersonnel for possible insights onperformance deficiencies and opera-

tional problems. This information,along with service data, was thenassembled into a file.

• Process data – including head andflow requirements – and physicaldata was then collected on the fluid

PUMP EVALUATION SUMMARY PUMP TYPE___________SISTER PUMPS_________

UNIT___________NO.__________TRI NO. PU__________________MODEL______________________DATE EVALUATED___/___/___

# OF NOTICES ________________LAST NOTICE ________________PPM ________________DATE ________________MAINTENANCE COST OVERLAST 7 YEARS _______________$# OF SEAL FAILURES __________PERIOD_________ MTBF________

IS THE SEAL CURRENTLYMEETING THE 1000 PPM LIMIT?❑ YES ❑ NO

HYDRAULIC PERFORMANCE

HEAD

FLOW

NPSHA/R

VAP. PRES.

TEMP.

S.G.

ORIGINAL EXISTING %BEP ADEQUATE?

❑ YES

❑ NOYES

PRELIMINARY SEAL SELECTION❑ SINGLE ❑ TANDEM

NO

NO

POSSIBLE REVAMP?❑ YES ❑ NO

YES

AFTER 1960 DESIGN?❑ YES ❑ NO

DESIGNED FOR SEALS?❑ YES ❑ NO

SHAFT DEFLECTION AT SEAL< 0.002 INCHES?

❑ YES ❑ NO

STUFFING BOX VENT?❑ YES ❑ NO

SUFFICIENT STUFFING BOX AREA FOR TANDEM CARTRIDGE?

❑ YES ❑ NO

BEARING HOUSINGACCEPTABLE?

❑ YES ❑ NO

STEEL BEARING HOUSING❑ YES ❑ NO

ALL 7 YES

SEAL UPGRADE

SEAL SYSTEM $___________

FIELD MODIFICATIONS $___________

PIPING $___________

INSTRUMENTATION $___________

REMOVE/INSTALL $___________

PUMP MODIFICATION $___________(INCL CPLG)

_______________________ $___________

TOTAL $___________

REMARKS

CASING DESIGN❑ RADIAL ❑ AXIAL

YES

RADIAL

RESCHEDULE

____/____/____

PUMP CASING MATERIAL❑ CS ❑ SS ❑ CI ❑ NI

CI or NI

CS or SS

AXIAL

AT LEAST 1 NO

REVAMP

❑ RETROFIT BACK PULLOUT ❑ OTHER

SEAL SYSTEM $___________

FIELD MODIFICATIONS $___________

PIPING $___________

INSTRUMENTATION $___________

REMOVE/INSTALL $___________

PUMP MODIFICATION $___________(INCL CPLG)

_______________________ $___________

TOTAL $___________

NEW PUMP

SEAL SYSTEM $___________

FIELD MODIFICATIONS $___________

PIPING $___________

INSTRUMENTATION $___________

REMOVE/INSTALL $___________

PUMP MODIFICATION $___________(INCL CPLG)

_______________________ $___________

TOTAL $___________

Figure 5. Pump evaluation summary

218 The Pump Handbook Series

Page 218: Centrifugal Pumps Handbook

being pumped. The information,which included vapor pressure andsolids concentration, was then sum-marized on a Pump Evaluation Sum-mary form (Figure 5) and used todetermine the best fix based onpump type, emissions, maintenancehistory and performance data.

Selection – The pump and sealsalliance consultants submitted pro-posals for the agreed upon upgrades.Attached to each was a seal proposaladdendum (Figure 6) that provideddesign details for construction. Theseals consultant also prepared a newseal order checklist (Figure 7) to fur-ther define the construction details.When field measurements wererequired, the pump was taken out ofservice and checked to ensure thatall the components fit precisely.

Installation & Startup – After theinstallation was completed, a QA/QCevaluation was made, and the PumpCommissioning Check List (Figure 8)was signed by the project representa-tive and the operator prior to startup. The seal vendor usually witnessedthe startup and recorded initial emis-sions levels.

Living Program Maintenance – As part

of the long range compliance strate-gy, data is still being collected on allVOC equipment to determine futuredirection. Some areas where thealliances are now concentrating their efforts include:

• Providing on-going training formaintenance and operations person-nel featuring detailed information onseal installation and operation. It isexpected that this training will great-ly increase the MTBR and reduce lifecycle costs of pumping systemsthroughout the refinery.

• Continuing to develop records onseal life, failure analysis and lifecycle costs, with focus on solutions tobad actor pump/seal systems.

• Incorporating seal and pump partsand repair services under a singlemanufacturer for each.

Conclusions andRecommendations

Accomplishments

• All of the 102 pumps modified in the project beginning in 1991 met1993 emission limits.

• More than 60% of these pumps hadinitial emissions levels of 1000 ppmor more, and the MTBR initiallyaveraged 8 months; after retrofittingthe MTBR has increased to an aver-age of 16 months.

• The alliances established criteriafor cost effective procurement ofpumps and seals, and they providedaccessibility to the most current tech-nology resources.

• An engineering standard for fur-ther VOC pump upgrades and arepair and installation standard forpumps and seals was put in place.

A key contributor to this successwas the ability of those involved to

FUGITIVE HYDROCARBONEMISSIONS PROJECT PUMP

COMMISSIONING CHECK LIST

PUMP NO.:_______________ DATE:__________________

MACHINISTS: Name:____________________________________

1. COUPLING GUARD SECURE _____

2. HOLD DOWN BOLTS INSTALLED PUMP _____

MOTOR _____

3. FLANGES PROPERLY MADE-UP _____

4. JACKING BOLTS BACKED OFF _____

5. SEAL DRIVE COLLAR BOLTS TIGHTENED _____

TORQUE: ___________________

6. SEAL SETTING PLATES ROTATED AWAY FROM COLLAR _____

COMMISSIONING ENGINEER: Name:___________________

1. PROPER OIL LEVEL PUMP _____MOTOR _____

2. PRESSURE GAUGES INSTALLED/ORIENTED _____3. PIPE PLUGS INSTALLED _____4. GASKETS INSTALLED _____5. SMALL BORE PIPING SUPPORTED _____6. NOMALLY-CLOSED VALVES CLOSED _____7. ORIFICE PLATE INSTALLED WITH INSCRIBED TAB _____8. COOLING WATER FLOWING9. VENTING PROCEDURE SIGN POSTED

AND VALVES TAGGED _____10. AREA CLEANED UP _____11. LOCKS/TAGS REMOVED _____12. VENT PROCEDURE DELIVERED _____13. SCREWED PIPING LEAK TESTED _____

Figure 8. Pump commissioning checklist

IF THERE ARE ANY PROBLEMS WITH THISPUMP AFTER COMMISSIONING CALL STEVE

ROSSI AT EXT.3263

SEAL PROPOSAL ADDENDUM

PUMP#/UNIT: PU.... /PUMP REPAIR TYPE:SEAL:PRODUCT:TEMPERATURE: FSUCTION PRESSURE: PSIGVAPOR PRESSURE: PSIADISCHARGE PRESSURE: PSIGSPECIFIC GRAVITY:API PIPING PLAN:SUCTION RETURN?:QUENCH?:PROPOSAL#:TOSCO CATALOG#:PUMP TYPE:NUMBER OF BOXES:COMMENTS:

SIGN-OFF: STEVE ROSSI_________GIL TIGNO__________

Figure 6. Seal proposal addendum

Figure 7. New seal order checklist

NEW SEAL ORDER CHECKLIST -

FOR OVERHAUL PUMPS

PUMP NUMBER: _______________

FIELD MEASUREMENTS

1) Physically verify: [ ]Shaft diameter___________ Bolt circle_____________Stud size_________________ First Obstr____________Gage Ring dist____________ Bolt Orientn__________

2) Suct Press___________ Disch Press ___________ [ ]

3) Rotation from driver end - CW/CCW [ ]

4) Temperature ___________

5) Make a diagram of the bearing web and the exis- [ ]ting seal piping. Where can new seal piping be located?

6) Note cage ring tap location(s). Can the seal box be [ ]vented through the cage ring taps?

7) Is the existing seal the same model as indicated by [ ]the files?

8) What is the O.D. of the current seal gland. [ ]

EVALUATIONS1) Verify the vapor pressure if possible. Make sure [ ]

that the box pressure is sufficient to keep adequate vapor suppression.

2) Design the seal flush piping system including [ ]orifice sizing and throat bushing clearance in order to get the required flow and vapor suppression.

3) Make sure there is adequate room for an O-Ring [ ]groove and multi-port injection between the box bore and the inside the stud holes. (especially important if box is to be bored)

4) Verify that the seal selected will fit. OK any box [ ]boring that will be required with Tosco and pump manufacturer.

The Pump Handbook Series 219

Page 219: Centrifugal Pumps Handbook

view the solution as an overall sys-tem of modifications. Pre-existingconditions such as pipe strain, unsta-ble foundations and misalignmentwere corrected to eliminate vibra-tion, stresses and distortions. As anadded benefit, the reliability andsafety of the equipment improved,thus lowering equipment life cyclecosts.

Correlation of Bad Actors to EmissionsCompliance

From 1990 to 1992 numerous in-kind maintenance repairs were madeon equipment in response to emis-sions violations. Most of theserepairs lasted only 3 to 6 monthsbefore another violation notice wasreceived. The upgrades undertakenhave in many cases doubled ortripled the time between emissionsfailures and eliminated chronic relia-bility problems.

Lessons Learned

There are three primary causesfor premature seal failures and/orexcessive vapor emissions fromupgraded pumps:

• installation errors• changes in the chemical composi-tion of the pumpage • operational and hydraulic problems(such as dry running and cavitation)

The first problem is the mostcontrollable. The others are morechallenging and require continuouseducation and training.

In addition to initial equipmentinstallation, an improved focus onequipment reliability through trou-bleshooting to resolve premature fail-ures is needed. This is expected totake the form of additional trainingfor both maintenance and operatingpersonnel, revision of operating pro-cedures, and continuous measure-ment of MTBF and life cycle costs.

A skilled team of dedicatedexperts can rebuild a pump perfectlyand still fail to achieve the finalobjective if the system is not startedup and operated properly. A numberof details must be attended to inorder to achieve success:

• Include process operators in theinstallation process. Have the pro-

duction department assign responsi-ble operators to the start-up team.Communicate their responsibilitiesfor a proper start-up and continuedoperation. Then conduct training onany special requirements of the sealsystem. Use seal and pump partnersto develop materials and providetraining.

• Develop pre-startup checklists thatinclude the following procedures:

a. Steam, flush and purge the pump casing prior to introducingproduct. (Minimize the time spent doing this to avoid contam-ination and overheating in the seal chamber.)

b. Prepare for hot alignment checks (P.T. > 300ºF and steam turbine driven pumps).

c. Review existing pump start-up instructions.

• Prior to starting a pump, gatherresponsible core team memberstogether, including alliance partners.Review the start-up procedure andthe duties of each team member.Develop a start-up checklist thatincorporates the following informa-tion:

a. Pump start-up procedures (including venting of all air and vapors from the seal chamber prior to and during start-up.)

b. Expected normal, minimum and maximum operating para-meters (flow, temperature, pres-sure, viscosity, cooling, etc.)

c. Performance parameters, including suction and discharge pressures, flow temperatures, suction strainer differential pres-sure and so on.

d. Program for continuous moni-toring after start-up.

e. Troubleshooting guidelines foroperators and mechanics.

• Pump and seal alliance partnersshould be full participants with usersin the successful commissioning andoperation of retrofit pumps. As stat-ed, they conditionally guaranteetheir equipment if all repair, installa-

tion and start-up conditions are met.For this project, the seal alliancepartner guaranteed that fugitiveemissions levels would not exceedBAAQMD limits for three years ofcontinuous operation.

• The ability to exercise a warranty isdependent on good documentation.Post start-up documentation require-ments must be agreed to withalliance partners as part of the initialparameters of the arrangement. As aminimum, the following data shouldbe collected:

Fugitive emissions levels – Initial-ly, the project team collected thisdata monthly until levels stabi-lized, at which time the monitor-ing was turned over to the con-tractor responsible for collecting quarterly compliance data.

Vibration data – This information is also taken more frequently in the beginning, to catch infant mortality-type failures. When readings stabilize, then routine (documented) monitoring can resume.

The importance of documenta-tion can not be overstated. Properdocumentation [API Standard 682,Shaft Sealing Systems for Centrifugaland Rotary Pumps, First Edition,October 1994] is required throughoutthe entire process from start to fin-ish. The minimum requirements arelisted below:

Seals

1. Completed API Standard 682 datasheets2. Cross sectional drawing of all seals(modified typical)3. Schematic of any auxiliary system(or systems) including utility require-ments4. Electrical and instrumentationschematics and arrangement/connec-tions5. Seal manufacturer qualificationtest results, if specified6. Detailed cross sectional drawingsof all seals (specific, not typical)7. Detailed drawing of barrier/bufferfluid reservoir (if included)8. Detailed bill of materials on allseals and auxiliaries

220 The Pump Handbook Series

Page 220: Centrifugal Pumps Handbook

9. Material safety data sheets on allpaints, preservatives, chemicals andspecial barrier/buffer fluids10. Installation, operation and main-tenance manuals11. Pre and post start-up checklists12. Routine performance monitoringdata sheets

Pumps

1. Completed API Standard 610 datasheets2. As-found and as-built specifica-tions (rebuilt pumps)3. Pump manufacturer performancetest results, if specified4. Detailed cross sectional drawingsof all pumps (specific, not typical)5. Detailed bill of materials on allpumps and auxiliaries.6. Material safety data sheets on allpaints, preservatives and chemicals7. Installation, operation and mainte-nance manuals8. Installation checklists9. Pre and post start-up checklists10. Routine performance monitoringdata sheets

On-going ProgramPerformance

In closing, the importance ofcontinuing to work within the

alliance partnerships cannot beoveremphasized. These relationshipsrequire constant nurturing and atten-tion. Resistance to using the alliancewill be an ongoing issue for the coreteam members. The alliance mustconstantly review the performanceof the partnership itself and compli-ance with stated goals and objectives.A formal and periodic review processshould be formulated.

Additionally, long term issuessuch as how to provide continuousimprovement (CI) to the alliance rela-tionship are important. There aremany facets to CI, but it typicallyinvolves empowering employees topursue improvements actively. It alsomeans providing technical support atthe front line, rigorous root causefailure analysis and use of advancedanalytical techniques. Along with CIis the need to maintain the new wayof doing business. This includes pur-chasing quality spare parts, main-taining quality control andstandardization, doing meticulouspump and seal overhauls, developingconsistent, detailed documentation,and retaining a highly skilled andmotivated work force.■

Stephen Rossi is Principal Engineerfor Rotating Machinery with Tosco

Refining Company, a Division of ToscoCorporation. Steve re-joined Tosco todirect the engineering effort for the com-pany’s Fugitive Hydrocarbon EmissionsProject in 1992. Previously he spenteight years at Chevron as a mechanicalequipment consultant on pumps, tur-bines, engines and compressors.

John B. Cary consults in the area ofreliability improvement and is currentlya Manager in the Maintenance Servicesunit for ERIN Engineering andResearch in Walnut Creek, CA. Mr.Cary has more than 20 years of experi-ence in the hydrocarbon processing andpetrochemical industry. Prior to joiningERIN, he was a Reliability Superinten-dent for Tosco.

Editor’s Note: This article has beenreproduced with permission of the Tur-bomachinery Laboratory. Edited fromProceedings of the Fourteenth Interna-tional Pump Users Symposium, Turbo-machinery Laboratory, Texas A&MUniversity, College Station, TX, pp. 25-34, Copyright 1997.

Acknowledgment: The authorswould like to thank Gil Tigno of theTosco Refining Company for his contri-butions to the success of this program.

The Pump Handbook Series 221

Page 221: Centrifugal Pumps Handbook

Evaluating SeallessCentrifugal Pump

Design & PerformanceCaught in the sealed versus sealless crossfire? This article sheds light on the appraisal

process so you can determine the best fit for your service.

By Dave Carr

Sealless pumps have been avail-able for more than 50 years,yet it has been only during thelast decade and a half that they

have gained popularity in NorthAmerica. Sealless pumps are usedmuch more often in the Europeanand Japanese markets. It is not coin-cidental that North America’s inter-est in this category of pumpscoincides with legislation aimed atprotecting the work place and thenatural environment throughincreased safety and minimization offugitive emissions.

Many reviews have previouslybeen published regarding the twodrive technologies that dominate thisclass of sealless pump – cannedmotor and magnetic couplings. Para-mount to both designs is the fact thatno dynamic shaft seal is required tocontain the pumped fluid. Rather, astationary containment shroud isused to isolate the pumpage from theambient environment. This is possi-ble because power is transmittedacross the shroud through magneticlines of flux that induce rotation ofthe impeller shaft. The significanceof this design feature is the fact thatsealless pumps can negate leakagethat is a fundamental byproduct ofthe typical mechanical face seal. Thecorresponding benefits to personneland the environment are obvious.

The chemical processing indus-try (CPI) is a leader in the adoption of

sealless centrifugal process pumpsand a model for their use in NorthAmerica. The CPI is currently expe-riencing a heightened awareness ofvolatile organic compound (VOC)and volatile hazardous air pollutant(VHAP) emissions as a result ofgrowing regulatory requirements.Today’s environmental regulationsalso have related monitoring andreporting costs in association withthe 1990 Clean Air Act. Seallesspumps fit within the broadest exemp-tion from emission instrumentinspections due to their compliancewith the without an externally actuat-ed shaft penetrating the pump housingdefinition [Ref.1]. These costs varyfrom plant to plant but, in aggregate,can account for a substantial bud-getary allowance with equipmentthat is pumping volatile liquids. Like-wise, they represent an input thatshould be included within an assess-ment of “total evaluated costs” (TEC)– i.e., all costs accrued from inquirythrough the life of the pump.

Numerous studies show thatshaft seals are the principal cause offailures in chemical process pumps.One such study [Ref. 2], of more than1250 operating pumps (and morethan 2500 installed pumps), indicatesthat approximately 50% of therecorded primary failure causes werethe result of leaking seals. Addition-ally, officials for a major chemicalmanufacturing plant [Ref. 3] have

concluded that their pump repairsare 30 to 60% less costly/frequentwith canned motor and magnetic dri-ven sealless pumps than with chemi-cal duty pumps equipped withmultiple mechanical seals. It is,therefore, important to recognize therole that maintenance/operatingavailability should play in the pumpevaluation process. When factoredinto the TEC equation, it has oftenbeen cited as the component thatsways a decision toward a seallessoffering.

Consequently, the need for asuccinct “sealed versus seallesspump” decision process has sur-faced. The following discussionreviews a logic stream that can beused to assist in this exercise. It isreasonable to expect that the out-come of any analysis model can notbe considered absolute since cus-tomer-specific application and expe-rience-sensitive inputs are oftenrequired to make an informed equip-ment selection. The basic tactic,however, can be used to determinewhen a sealless design is a viablechoice.

Sealless HydraulicCapabilities

Performance capabilities to 2000gpm and 750 feet TDH (though notnecessarily simultaneous) are repre-sentative of widespread seallesspump experience and coincide with

CENTRIFUGAL PUMPS

HANDBOOK

222 The Pump Handbook Series

Page 222: Centrifugal Pumps Handbook

single suction, single stage, centrifu-gal pump capabilities. Productiondesigns are available to furtherextend this region, but field experi-ence has been typically limited tohead and flow combinations that fitwithin 200 horsepower. Require-ments outside of this range will likelymandate the use of larger singlestage, multistage or high speed cen-trifugal pumps or positive displace-ment designs. The vast majority ofsealless manufacturers’ equipment isbuilt around 150 and 300 poundANSI flange ratings, which are con-sistent with the needs of a typicalchemical processing plant.

Once it has been determinedthat your service conditions roughlyfit into the range of today’s seallesscentrifugal pumps, an appraisalprocess to determine the best fit canbegin. Pump users are particularlychallenged when operating with liq-uids that are characterized by one ormore of the volatile, toxic, flammableor corrosive definitions. In the CPI,this represents a broad range of flu-ids that include acids, alkalis, salts,esters, hydrocarbons, monomers/polymers, alcohols, ethers, halo-genides, nitrogen/sulfur compoundsand even some extreme water condi-tions. Some characteristics commonto these broad definitions are apropensity for the fluids to be unsta-ble, poisonous, noxious, dangerous,destructive or chemically reactivewith air. They may also precipitatecomponents, solidify easily, or needto be handled at extreme tempera-tures. Sealless pumps are well suitedfor these applications, but some basicapplication criteria need to be estab-lished to ensure a troublefree instal-lation.

Look at the LiquidA qualification of the pumped

liquid is the first major decision pointwhen considering the use of a seal-less pump. Early applications wererelegated to services addressing aneed to meet emissions limits forvolatile fluids (VHAP/VOC) or wheremechanical seals had proven ineffec-tive. Often those applications wereassociated with situations in whichfluid volatility was a significant appli-cation parameter. It should immedi-

ately be recognized, however, that asimple change from sealed to seallessoptions is not appropriate without acorresponding analysis of thestrengths and suitability of eachpump. By design, a mechanical sealface requires a liquid film to act as acooling and lubricating mediumbetween the stationary and rotatingmembers. Corresponding leakageacross those faces, be it ever soslight, is a necessary consequence topromote acceptable seal life. This sit-uation is complicated with volatilefluids having vapor pressure charac-teristics that result in a change fromliquid to vapor states at atmosphericpressure. Consequently, leakage cancause liquid mechanical seals to be asource of fugitive emissions evenwhen operating properly.

A hybrid category worth specifi-cally noting is heat transfer fluids(HTF), which can exhibit one ormore of the above mentioned attrib-utes. Many users believe that HTFapplications represent the singlemost prevalent application for seal-less pumps in the CPI today sincethose services transcend safety, envi-ronmental and maintenance issues.Any of the three criteria by itself canbe used to justify a sealless pumppurchase decision and, collectively,they represent a prime opportunityto exploit the leak-free nature of seal-less pumps.

Only sealless and multiplemechanical seal (either liquid or gasconfigurations) pump designs shouldbe considered to meet demands ofliquids in the volatile, toxic, flamma-

ble, corrosive or HTF categories.Applications that fall outside of thoseparameters can reasonably be con-trolled with single mechanicallysealed pumps as an alternative tomore sophisticated designs.

Frigid or Fervid?Extreme temperature conditions

represent good opportunities for theuse of sealless pumps. This belief isattributable to the fact that such con-ditions present a difficult situationwith regard to maintaining the neces-sary seal face integrity that comple-ments low leakage. High values arecommonly associated with referenceto “extreme” temperature conditions.It is generally accepted that a “hightemperature” definition refers tooperating levels greater than approx-imately 250ºF (121ºC), and that tem-peratures above this threshold act asa handicap to the operational ease,simplicity and effectiveness of liquidsealing. The development of highstrength magnets and insulationmaterials has yielded sealless designscapable of withstanding tempera-tures as high as 750ºF (399ºC) with-out the use of auxiliary cooling.Current technology limits cooled ver-sions to as high as an 850ºF (454ºC)rating, but these limits are continual-ly being challenged by new productdesigners intent on extending appli-cation limits.

Cryogenic extremes, however,are often disregarded when consider-ing the use of sealless pumps. Mostsealless pump manufacturers ratetheir standard equipment for levels

Liquid Types: hazardous, or volatile vapors, toxic, flammable, corrosive heat transfer fluidFlows: <2000 gpm <750 ft TDH

Field experience typically limited to head/ flow combinations that fit within about 200 hp

Temperatures: <750°F (without auxiliary cooling)<850°F (with auxiliary cooling)>-140°F (specialized designs to -260°F)

Viscosity: >0.15 cP<200 cP(Note: Lower viscosities may require bearing design changes to support shaft)

Solids Content: <150 Micron3-5 wt %

SEALLESS CENTRIFUGAL OPERATING CHECK LIST

The Pump Handbook Series 223

Page 223: Centrifugal Pumps Handbook

to -140ºF (-96ºC), and specializeddesigns are available for tempera-tures approaching -260ºF (-162ºC). Aside benefit to a canned motorpump’s operation with cold liquids isthe fact that increased cooling capac-ity can increase the power transmis-sion capability of a particular framesize. The basic advantages that at-tract pump users to sealless designsfor high temperature services areequally pertinent for low tempera-tures.

The pumped liquid’s sensitivityto temperature changes must also beconsidered to complement its dis-creet temperature evaluation. Thisincludes liquids that have peculiarmelting and/or freezing points, e.g.MDI, TDI and many acids, and onesthat experience polymerization orcrystallization with accompanyingtemperature changes, such as formal-dehyde. Some liquids (caustics, forexample) experience crystallizationwhen they come in contact with air.This can occur as the liquid leaksacross a conventional seal face orwhen air is drawn into the pump –e.g., in a system where the pump’ssuction is less than atmospheric pres-sure. These circumstances are trou-blesome to mechanical seals sincethe crystal residue can be abrasive tothe sealing faces and inhibits faceadjustments with accumulation inthe secondary area. Also, in vacuumapplications process disruptions canoccur with the introduction of air.

Sealless pumps address theseproblems by nature of the pumpedliquid’s absolute isolation from theambient environment and the abilityto add or subtract heat from the sys-tem. Jackets are available for drivesections, bearing housings and pumpcases to meet minimum or maximumtemperature constraints of a particu-lar liquid. For the successful applica-tion of pumps in such services,however, the pump user must dis-close a liquid’s temperature sensitiv-ity characteristics. It is imperative aswell to recognize that unobstructedflow paths, i.e., without the accumu-lation of solids, crystals or polymers,are essential to ensure lubricationand heat dissipation within the drivesection. If a heat medium is not avail-able for liquids that have the propen-

sity to polymerize, crystallize orfreeze, an externally flushed seallessor a multiple mechanical sealedpump should be considered.

Thick or Thin?Centrifugal pumps are generally

applied to relatively low viscosity liq-uids, and hydraulic-end viscosity cor-rections are the same for a sealed orsealless design. When high viscosityliquids are passed through the drivesection, however, increased parasiticlosses occur. The Hydraulic Institutespecifies that an external flush fluidshould be used when the viscosityexceeds 200 centipoise [Ref. 4]. Anysimplified evaluation of viscosity’simpact toward parasitic losses mustrecognize the fact that smaller/slow-er speed designs are much less af-fected by viscosity changes. Thecorrection factor corresponding tothe drive section’s coolant/lubricantflow is an exponential step function,rather than a linear relationship,therefore parasitic losses must bewell understood. Special designs areavailable to extend a sealless pump’sviscous handling capabilities, but theearlier comment on the value of spe-cific experience, be it the manufac-turer’s or user’s, is again appropriate.

On the opposite extreme, manyliquids described in the “Look at theLiquid” section of this article haveviscosities that are less than that ofwater and can benefit from the use ofsealless pumps. Included within thisgroup are liquids such as HF acid,ammonia, fluorocarbon refrigerants,hot water, heat transfer fluids andhydrocarbons, all of which are regu-larly used in many CPI processes. Itshould be recognized that there arelong-standing debates about theacceptability of such liquids withproduct lubricated bearings. Experi-ence has shown, however, that theycan conservatively be used, withwell designed bearing systems, downto a 0.15 centipoise level, and somemanufacturers have exhibited experi-ence at even lower values. Properlydesigned product lubricated bearingsystems result in legitimate seallesspump alternatives to multiplemechanical seals for low viscosityservices.

A liquid’s contamination with

solids is another critical concern inthe appraisal of a sealless pump’sviability. Various manufacturers’ lit-erature specifies a range of 2 to 6%by weight. A 3 to 5% guideline is aslight compromise from the maxi-mum published range, but it accu-rately addresses the capabilities ofmost manufacturers and offers a rea-sonable rule of thumb. Clarificationwith regard to the size of the solids,however, is required since seallesspumps have small passages withinthe drive section circuit. A 150micron judgment limit allowsapproximately a 4:1 safety factor tothe clearance between the rotor andcontainment shroud with worst casedesigns. Invariably, one of the small-est clearance locations is at the shaft(sleeve) to journal bearing. Thisclearance typically ranges between0.001 and 0.0035 inches (radial) formany manufacturers. Therefore, fur-ther analysis is required to predictthe impact of the concentration, sizeand abrasiveness of solids withregard to the materials used for thebearings and other components ofthe drive section.

Pump applications that fail topass the critical questions discussedwithin the “liquid type” sections maybe more economically serviced by asingle mechanically sealed design.Extreme temperature and pressureconditions often prove to be anexception to the single seal decisionand present a competitive opportuni-ty for the sealless pump. There aretimes when the leak-free nature ofsealless pumps is desirable, but thetemperature sensitivity, viscosity andsolids content characteristics of thepumped liquid converge to demandan examination of a barrier liquid’sviability.

The reasonableness of using abarrier fluid to provide the drive sec-tion with a higher quality coolant/lubricant is analogous to clean liq-uids that would support mechanicalseals (API-32/52/53 seal supportplans). Disparities are generally theresult of supply and consumptiondifferences between the sealed andsealless configurations. Seallesspump manufacturers have becomeextremely innovative with flushdesigns for sealless drive sections,

224 The Pump Handbook Series

Page 224: Centrifugal Pumps Handbook

which effectively isolate the pump-age from that area. If a barrier fluidanalysis finds that the seallessapproach is not practical, the choicedefaults to multiple mechanicalseals, which will still require anancillary support system.

The Internal Flow CircuitMany of today’s sealless pump

applications are with liquids exhibit-ing good thermal stability. When thatis the case, the basic design under-standing of all manufacturers willensure stable pump operations.Volatile liquids, however, are anapplication field in which additionalscrutiny is required for operatingsuccess. We have previously dis-cussed the fact that a sealless pump’sinternal circulation system is used tosupply the bearings with lubricantand to dissipate the heat generated inthe drive section. The correspondingflow paths are demonstrated in Fig-ure 1 for a typical magnetic drivenpump.

Inadequate flow and/or pressurewill ultimately result in distress tothe bearings and, if unchecked, leadto ultimate failures. Active participa-tion in the equipment selection andoperation processes will minimizethese problems. The sealless decisionmust, therefore, include a thoroughengineering analysis for such ser-vices so as not to trade a seal prob-lem for one with a sealless pump’sdrive section.

Two critical liquid parametersthat must be understood are vaporpressure and specific heat. The pumpdesigner requires vapor pressureinformation at the rated temperatureand at some level greater than that tounderstand the fluid’s state as a

result of pressure andtemperature changes ofthe drive section’s flow.Normally, vapor pres-sures at rated tempera-ture plus 10ºF (5.6ºC)and 20ºF (11.1ºC) fromsuction will give thepump manufacturersufficient information tomake an informedanalysis. A fluid’s spe-cific heat value is criti-cal since most

manufacturers’ temperature riseanalyses are derived from testingwith water, which can absorb consid-erably more heat than many of theliquids in today’s critical processes.

A thorough understanding ofcoolant flow, pressure and tempera-ture must be demonstrated to sup-port troublefree pump operation inthe field. Pumps operating on volatileliquids typically exhibit minimumflow constraints that are dictated bythermal, rather than mechanical,limits. One of Murphy’s laws is thatpumps never operate at their designpoint. Therefore, the pump suppliermust also conduct his analysis at off-design points to establish a recom-mended operating “window” – i.e., aminimum to maximum flow regime– to give good operating flexibility.There are many successful seallesspump applications with steep vaporpressure liquids attesting to the factthat reputable manufacturers possessthe ability to meet these demandingservices. Ammonia, isobutane andpropane are examples.

The effort expended to conductthis flow circuit analysis will ensurethat the drive section is continuallysupplied with a liquid and not a gas.This situation is analogous to the top-ic of dry running a pump. There area number of design means that canbe used to minimize the effects of aloss of pumped liquid, includingimpellers with special or no wearrings, impregnated or coated bearingmaterials and nonmetallic shrouds.As with any pump, whether sealedor sealless, the goal should be to pre-vent or at least to minimize theoccurrence of a pump being operatedwithout liquid.

The successful application of

sealless pumps on numerous loadingand unloading services is one exam-ple of the viability of the designunder dry running circumstances.Success is conditional, however,upon the minimization of the dry runduration and having sufficient instru-mentation to protect the machine asconfigured. With a sealless pump, itmust be recognized that the driveand pump sections may run dry andequal care should be exercised toprevent either occurrence. Recenttechnological advances have beenmade with non-invasive electricaldevices that can sense the fluid con-dition – liquid, gas or two phase state– within the sealless pump’s drivesection, thereby ensuring reliablepump operation.

InstrumentationToo often sealless manufacturers

are confronted with installations inwhich no precautions have beenmade, supplied instrumentation hasnot been installed or excessive instru-ments (which yield unnecessarycomplications for field personnel)have been employed. A balancedinstrumentation plan is recommend-ed to preclude the possibility of theseoccurrences.

Flow measurement is the surestway to guard against a loss of pumpflow, but it is typically not availablein most chemical plant installations.A pseudo measurement of flow is achange in the driver’s power require-ment. For electric motors greaterthan approximately 10 horsepower, acurrent sensing device works well,but smaller motors may be betterprotected with watt meters. This pro-tection should not, however, be usedas a minimum flow protectiondevice! Power or current sensingdoes not have the sensitivity to act inthat manner but can readily indicatethe onset of a “no-flow” condition.

Pump flow and power instru-ments will not indicate a vaporiza-tion state in the drive section that hasbeen caused by blocked flow pas-sages or insufficient cooling. A tem-perature detection device, such as anRTD or thermocouple,is commonlyused to indicate that condition andcan be positioned to sense the tem-perature of the containment shroud.

Figure 1.

The Pump Handbook Series 225

Page 225: Centrifugal Pumps Handbook

It is generally accepted that a nomi-nal 20ºF (11.1ºC) setting, greaterthan the calculated temperature rise,will protect against a dry run occur-rence. Further, insufficient coolingwill cause a CMP to experience arapid increase in the stator windingtemperature, a fact that highlightsthe need to always wire its ther-mostats.

To meet electrical safety stan-dards, Europeans have used the sci-ence of ultrasonics to ensure a “wet”condition within the drive section oftheir canned motor pumps. Innova-tive adaptations to this technologyare finding their way into seallesspump designs in North America toguard against overheating a motor ormagnet assembly. Similarly, instru-ments are now available to indicate acanned motor pump’s direction ofrotation without physically seeingthe shaft turn. This advancementaddresses a long-standing complaintby users resulting from instructionsto observe a dry run bump start toguarantee proper motor wiring.

A number of manufacturers nowoffer bearing wear detectors for usewith carbon bearings. These instru-ments can be either electrical ormechanical and do a good job ofalerting the user to an impendingminimum material condition. Conse-quently, routine maintenance can bescheduled in a manner consistentwith production needs. Growingcompetition in today’s sealless pump

marketplace has prompted manufac-turers to take a more active role inthe application of pumps, and theircounsel should be sought whendeveloping a sealless pump instru-mentation plan for the first time.

This article has attempted todemonstrate that sealless centrifugalpumps are a valuable asset in thearsenal of today’s pump users. Thefact that they completely eliminatefugitive emissions complements in-creasingly strict environmental andsafety objectives. To capitalize on asealless pump’s cost effectivenessand inherent reliability, however,applications should be well under-stood. Wisdom in this area resultsfrom a combination of complete cus-tomer input, a manufacturer’s thor-ough technical analysis/applicationexpertise and adequate protectiveinstrumentation. Worldwide experi-ence shows that these pumps shouldnot only be considered for new in-stallations. Many conversion oppor-tunities exist for the retrofit ofexisting, poorly performing pump in-stallations. Manufacturers’ standard-ization with ANSI dimensions facili-tates this process and makes retro-fitting a cost effective alternative toliving with leaking pumps.

There will continue to be a needfor mechanically sealed pumps thatoperate outside of the hydraulic ca-pabilities of sealless equipment. Move-ments toward a continuous improve-ment philosophy, increasing recogni-

tion of sealless pump maintenanceadvantages, continued developmentsin materials and significant stridestoward pricing parity can be expectedto further encourage the worldwideutilization of sealless technologies.■

References:1. Guidelines for Meeting Emission

Regulations for Rotating Machinerywith Mechanical Seals, STLE SpecialPublication SP-30 (1994)

2. Bloch, H.P. Practical Machin-ery Management for Process Plants,Improving Machinery Reliability, GulfPublishing Company (1982)

3. Fischer, K.B., Seifert, W., andVollmuller, H. Canned Motor andMagnetically Coupled Applications,Operations and Maintenance in aChemical Plant, Proceedings from theTenth International Pump UsersSymposium, Houston, Texas (1993)

4. American National Standard forSealless Centrifugal Pumps ANSI/HI5.1-5.6, Hydraulic Institute (1994)

David M. Carr is a Project Manag-er with Sundstrand Fluid Handling inArvada, Colorado and a frequent con-tributor to Pumps and Systems. He hasbeen with the company since 1980 andhas held various positions within thesales, marketing, product engineeringand product development departments.He is a graduate of Purdue Universitywith a B.S. degree in mechanical engi-neering technology and an M.S. degreein management.

226 The Pump Handbook Series

Page 226: Centrifugal Pumps Handbook

CENTRIFUGAL PUMPS

HANDBOOK

The Pump Handbook Series 227

Vertical Motor-UnderPumps Expand Their Range

By Pumps and Systems Staff with Gaylan Dow of Hayward Tyler, Inc., and Joe Campanelli of Air Products & Chemicals, Inc.

If you asked two individuals, onean experienced utility machineryengineer and the other an experi-enced process machinery engi-

neer, to describe an integral motorsealless pump (think canned motorpump for now) you probably will gettwo very different descriptions. Likethe proverbial elephant and the blindmen trying to understand it throughtouch, sealless pump technology isperceived differently at differentends of the marketplace. This articlehighlights the features of a type ofsealless pump utilized extensively inthe utility industry. It also explainshow many of these features are beingused with benefits in process envi-ronments.

Most specifying engineers andpump users are familiar with thesealless pump/motor combinationdesign of a canned motor/centrifugalpump of end suction configurationwith a single primary radial impellerthat does the work. It is mountedhorizontally on a bracket or base-plate, and, short of being compact indesign and without a coupling, itwould look similar to most processpumps. If you pressed harder, youmight learn that it uses a mechanismknown as “hydraulic balance” tolocate the impeller pretty much inthe center of the volute.

Another class of sealless integralmotor pumps has been used fordecades in the utility industry. Inaddition to boiler feed pumps foundin all conventional power plant boil-ers, many boilers are designed to usepumps that force the water to circu-late instead of relying on natural con-vection circulation. With thousands

of units installed worldwide, the ver-tical motor-under pump is one of theunsung success stories in the pumpindustry.

Proven Design for UtilityService

Boiler water circulation is ardu-ous duty, with pressures around3000 psig, temperatures of about 650 ºF and horsepower to 1600. Thepump that has evolved to suit thisduty is fabricated into and suspend-ed by the plant pipe work (Figure 1).In this setup the motor is slungbeneath the pump and is, in turn,supported by a hot neck, which ther-mally isolates the pump and motor.The whole assembly, not tied to anybaseplate or foundation, is free tomove with thermal expansion of theplant. In many instances the pumpswill move more than one foot fromcold start-up to hot running condi-tion. This movement representsrelief of piping strain which wouldnormally have been imposed on thepump through the nozzles.

Water at boiler pressure is circu-lated through the motor and heatexchanger by a secondary (auxiliary)impeller that is designed into thethrust bearing. Because of the largehorsepower involved, these units aredesigned to maximize the efficiencyof the motor. However, though theyare highly efficient for fluid-filledmotors, even a small fraction of alarge horsepower yields a significantamount of waste heat that needs tobe disposed of.

This precludes the use of acanned type approach with a linerthat separates the winding from the

fluid in the bore of the motor. Theheat generated would create unac-ceptably high temperatures in thewinding cavity.

Consequently, a design approachsimilar to that used in many types ofbore hole submersible pumps isused. The windings are coated with ahigh dielectric strength polymer andimmersed in the water itself. This istermed a WSU (wet stator unit) (Pho-to 1) as opposed to a DSU (dry statorunit/canned motor), and it ensuresthat a large surface area is cooled bythe internal water flow. Additionally,with the absence of an internal liner(or can), the rotor to stator gap isincreased, and a failure point (dam-age to the can) is removed.

Figure 1. The glandless recirculationpump is fabricated into and suspend -ed by the plant pipe work.

Page 227: Centrifugal Pumps Handbook

A special power feed-through isutilized. It forms both the terminalfor connection to the power supply(460V to 6.6 KV) and part of the pres-sure vessel of the motor.

Practical AdvantagesAdvantages of the motor-under

design include the following:

• The unit is continuously selfventing. Most conventional horizon-tal canned motor pumps with hotnecks and heat exchangers have thelatter located horizontally above themotor. On hot standby, this arrange-ment promotes thermal siphoningand keeps the motor insulation sys-tem within temperature limits. Thesetup will often require a ventingoperation from the high point (theheat exchanger) when initially filled.This operation is not required withthe motor mounted beneath thepump.

• The pump remains full untilintentionally drained. These are largepumps that are full of fluid and there-

fore extremely quiet. There havebeen reported incidents of boiler cir-culator pumps being left on while theboilers have been completelydrained. Although this is not recom-mended, having the motor (andtherefore bearings) surrounded byfluid minimizes the potential fordamage during dry pump running.

• The design utilizes a hydrody-namic thrust bearing to locate therotor axially. Because of the rotor’smassive weight on boiler circulators,generating and closely controlling thelarge thrusts necessary to accomplishaxial hydraulic balance is impracti-cal. Therefore, the pump is designedto thrust positively in one direction,and it incorporates thrust bearingscapable of accepting thrust loads inall cases. This provides the addedbenefit that the pump has a wide lat-itude of operation over the curvebecause there is no hydraulic balanceto upset.

• Radial bearing loads are re-duced by the weight of the rotor.Additionally, a multi-vaned diffuserin a concentric casing is sometimesused to further minimize radial loads.

• Suction conditions are opti-mized. Rather than mounting anelbow on the pump suction as is typ-ical of an end suction pump, a longer

run of straight pipe can come downdirectly into the pump suction.

• Design and construction costsare minimized for hot systems. Thepiping goes in, the piping goes out.Much like a motor operated valve,there is no need for long runs of pipe,numerous elbows or expansion jointsto relieve the piping strain on thepumps. This greatly simplifies theanalysis and design of the piping sys-tem. There is no foundation to design or pour to be done and no baseplateto design and fabricate. Also, there isno grouting or hot coupling align-ments.

• Heat tends to travel sidewaysor up. Locating the motor below thehot pump end minimizes the heatexchanger load and lifetime operat-ing costs.

• In addition to boiler circulationduty, this configuration is also used inthe utility industries in attemperatorspray pumps as well as in nuclearapplications such as reactor internalpumps and reactor water clean uppumps. To minimize the pipinginvolved with reactor internalpumps, the pump ends are actuallylocated directly in the reactor vessel.

Process VersatilityBecause the pump is continuous-

ly self venting, it eliminates an opera-

228 The Pump Handbook Series

Photo 1. Wet stator unit. A secondaryimpeller incorporated in the motorrotating element circulates waterthrough the motor cavity to the top ofthe heat exchanger and back to thelower part of the motor. This recircula -tion dissipates the heat generated bymotor electrical losses.

Photo 2. A 6000 psig/300ºF canned motor-under pump in service at a chemicalprocess plant. Large quantities of gas are injected and consumed as part of theprocess. This gas becomes entrained and the primary concern is loss of liquid inthe pump end. The process has upset conditions with rapid pressure spikes to6000 psig .

(photo courtesy of Hayward Tyler, Inc.)

Page 228: Centrifugal Pumps Handbook

tor function such as venting on start-up – always a good idea. Additional-ly, some process systems (such asheat transfer) generate non-condens-able vapors.

In other applications, gas that isinjected purposefully as part of theprocess is consumed, and more and more systems have gas injected atseals and as blankets. This gas tendsto collect through centrifugal forceabout the pump shaft and can makeits way back into the motor cavity.On a canned pump with a top mount-ed heat exchanger, this usuallyrequires the use of a level detectiondevice at the heat exchanger, as wellas operator intervention to vent thepump during operation. Even oncanned or other sealless pump appli-cations not requiring heat exchang-ers, the gas can centrifuge out aroundthe rotor and cause possible dry run-ning of the bearings.

In addition, because the pumpremains full of liquid until intention-ally drained, it is protected againstdry running, and the motor is imper-vious to large quantities of entrainedgas in the pump end. Typical applica-tions which this benefits include tankcar unloaders and batch processes inwhich tanks are completely evacuat-ed on purpose.

The sealless integral motorpump utilizes a hydrodynamic thrust

bearing to locate the rotor axially. Acanned motor pump relies on a set ofclearances about the impeller to bal-ance the rotor assembly hydraulical-ly and locate it axially. Theoretically,this could be accomplished in a verti-cal motor-under design by designingin just enough thrust to lift the rotorand locate it axially. In practice, how-ever, using a real thrust bearing hasits advantages. Positioning it in thebottom of the unit protects the pumpand motor assembly from systemupsets. Dry running, or the collectionof entrained gas behind the impeller,negates the lift generated by theimpeller. Thermal shock dramatical-ly changes the internal pump wearring clearances, which temporarilyand drastically alters the dynamics ofa thrust balance design. Also, historyhas shown that despite efforts to con-trol water chemistry, many boilersare loaded with oxides, which have atendency to wear clearances andblock balance holes, upsetting thethrust characteristics of the impeller.In various process applications rust,scale, catalysts or product solids canbe present, and these can create sim-ilar circumstances. Any of these con-ditions would cause a hydraulicallybalanced unit to lose lift or thrustupward enough to generate consider-able wear and damage to the pumpand eventual failure.

Radial bearing loads are reducedwith the motor-under pump. Thiscan be a real advantage for processesin which fluid viscosity is low. Andfor service in hot systems, thesepumps are cost effective to install.The same design constraints of tryingto bolt a massive hot piping system tograde through the pump are presentin process plants as well. The pipingwants to move, and it is better to letthe pump move with it.

Vertical motor-under cannedmotor pumps can also be attacheddirectly to a chemical reactor. Thissetup further reduces piping (andthus lowers costs) as well as mini-mizes the heat tracing requirements.

ConclusionA vertical motor-under canned

pump can be useful in applicationsrequiring zero leakage, including sit-uations in which large quantities of

gas are present or the pump runs dry.It can also provide reliable service forhot systems in which significant ther-mal strain can be imparted to thepump flanges, or on systems whichrun both hot and cold and alignmentis an issue.

If other service requirements canbe met, the motor under design canalso lower plant construction costsand be an advantage when floorspace is at a premium – such as onoil production platforms or in multi-level chemical processes. Last, it isalso useful in situations in which theminimization of piping is attractive,such as those requiring heat tracing,or for lethal services.■

Gaylan Dow is Sales & MarketingManager for Hayward Tyler, Inc.

The Pump Handbook Series 229

Photo 3. This canned motor-underpump at a chemical processing facility– an end suction confi guration – hasto contend with entrained gas andthermal shock.

(pho

to co

urtes

y of H

aywa

rd T

yler,

Inc.)

UpgradingBoiler WaterCirculation

ProcessBy Joe Campanelli, Air Products

and Chemicals Inc.

Process waste heat boiler cir-culating pumps have tradi-tionally been horizontalpumps of either end-suc-

tion, overhung impeller design (forvery low flowrates) or double-suc-tion, impeller-between-bearing,single-stage units for higher flowapplications. Boiler water circula-tion is a very demanding service,with the pump taking its suctiondirectly off the boiler steam drum.Due to the elevated suction pres-sure and temperature duty, me-chanical seal problems arefrequent, as are other maladiescommonly associated with highpressure/temperature pumps, in-

Page 229: Centrifugal Pumps Handbook

230 The Pump Handbook Series

cluding:

• Pump casing deflection. Thisis caused by thermally inducedpipe strain resulting in rotor rubs,seal leakage/failures, bearing fail-ures, casing cracks, casing splitlineleakage and piping cracks.

•Thermal stratification in hotstandby units. This results in rotorbows, seal leakage, internal rubson startup and rotor lockup.

• Seal leakage and seal mainte-nance problems. These are con-stant issues on boiler circulatingpumps – often requiring replace-ments and additional spares. Andcatastrophic seal failures, whilerare, do occur and can exposeoperating and maintenance per-sonnel to significant releases ofhot, high pressure water, whichimmediately flashes into a plumeof steam.

The recent introduction of ver-tical canned motor pumps into thisapplication has enabled our pro-cess pump users to capitalize onthe extensive experience gained byHayward-Tyler (HTI) in the utilityindustry. The HTI vertical cannedmotor pump offers a proven,robust mechanical design withoptimum materials of construc-tion. The application of these

pumps to process waste heat boilercirculation service has solved theproblems associated with conven-tional horizontal pumps and hasprovided significant benefits inpump reliability, maintenance andspare parts cost reductions.

Two recent waste heat boilerapplications that have utilized theHTI pumps with good results are detailed in the chart.

The pumps in both applica-tions have performed extremelywell to date and will be chosen for

this type of application on futureprojects. ■

Joe Campanelli is a LeadMechanical Engineer in the Produc-tion and Delivery Organization of AirProducts and Chemicals Inc. He has19 years of experience on a wide vari-ety of rotating equipment in theprocess industry and is currentlyresponsible for the safe, reliable andefficient operation of the machineryin 28 process plants.

Facility LaPorte, TX Pasadena, TX

Pump manufacturer HTI HTIPump type Dry stator Wet statorPump model 3 x 4-8/DSU 12 x 14-14/200 WSUCapacity - gpm 656 8052TDH - ft. 170 76Inlet Pressure - psig 745 620Discharge Pressure - psig 802 646Inlet Temperature - °F 510 489Speed - rpm 3490 1780Power - hp 72 200Pumps installed (4) 2 trains, 2x 100% (2) 2 x 100%On stream date Jan ‘96 Oct ‘96

Page 230: Centrifugal Pumps Handbook

Pump not working properly?Chances are something iswrong in the suction sector.To begin with, there are a

multitude of circumstances that cancause suction problems. To cover thissubject effectively and systematical-ly, you are being provided a formatfor checking existing operations aswell as information on properdesigns for new suction systems inthe planning stage.

The system’s suction arrange-ment must provide the energy tomove the liquid to the eye of therotating pump impeller. Until the liq-uid reaches the leading edge of theimpeller vane, the pump cannotimpart its energy to move the liquidonward.

One must establish a suctionrequirement for a pump that meetsor exceeds the system’s minimumsuction availability. It is also theuser’s responsibility to see that thesuction system provides and main-tains the stated condition at therequired flow and that the pump ismaintained in good operating condi-tion.

Existing OperationWhen a previously successful

pump-suction system fails, there arethree primary areas that need to belooked at for probable cause. Theyare the pump, the suction and theliquid being pumped. The most com-mon causes of failure in each sectorare:

Pumps • Leaky seal• Reduced speed • Pump modification• Wrong direction of rotation• Plugged impeller• Worn wearing rings

System Line • Leaky suction line

• Plugged line (including pump inlet)

• Sticky valve• Suction height increased • System line modification• Corrosion or product build-

up in suction line

Liquid • Increase in temperature• Drop in liquid level• Aeration due to vortex or

processing change• Viscosity change• Specific gravity change• Other physical property

changes

Once the problem is identified, itcan be corrected. In the case of a pro-cessing change, it provides aware-ness of potential problem areas tocheck new requirements against thecapabilities of the existing pump tosee if the unit remains adequate.

The eye-balling method of mea-suring cannot be relied upon todefine operating conditions effective-ly. Use proper instrumentation tomeasure pressure, flow, suction lifts,speeds and the temperature rise ofmotors.

When motor speeds are incor-rect, check connections and measurevoltage at the motor terminals.Remember, too, that flow and pres-sure readings should be taken inareas where stable accurate readingscan be obtained. This usuallyrequires 5 to 10 pipe diameters ofstraight piping after going through anunsettling section such as one thatincludes elbows, valves and reduc-ers. In addition to measurement, be aware of sounds different than those emanating during normal operations,as well as vibrations and surges.

Planning StageIt is during this phase that most

potential problem areas can best bedealt with. Begin with careful plan-ning and installation of the pumpsuction line. Design the line to pro-vide the shortest practical directroute utilizing large radius elbowswith an absolute minimum numberof fittings and valves.

Suction PipingAir pockets or high spots in a

pump suction line invariably causetrouble. Piping must be laid out so itprovides a continual rise or at least aperfect horizontal run without highspots from source of supply to thepump. For the same reason, aneccentric reducer instead of astraight taper should be used in ahorizontal suction line. Another wayto remove vapor trapped in a suctionline because of a high point is to ventthe sector back to the vapor space inthe supply vessel. If an air pocket isleft in the suction pipe when thepump is primed, it will often pumpproperly for a time and then lose itsprime or have its capacity greatlyreduced. The small pocket of airunder a partial vacuum conditionwill expand and greatly reduce theeffective flow cross-section of thepipe, thus starving the pump. Or airwill be drawn into the pump, result-ing in loss of prime. The suction pipeshould be submerged to a depth of atleast 3 feet when the water is at itslowest level.

Frequently, foot valves areinstalled on the end of suction pipefor convenience in priming. They should be sized to provide a flow area 50% greater than the pipe area.To protect the foot valve, pump andother equipment from being fouledby refuse such as sticks, rags, lightplastics and miscellaneous somewhatbuoyant solids, a screen strainer withan area at least three times that of thepipe area should be installed ahead

CENTRIFUGAL PUMPS

HANDBOOK

The Pump Handbook Series 231

Centrifugal Pump SuctionPump problems often begin on the suction side. Here is a guide for checking existing operations

as well as information on proper designs for new suction systems in the planning stage.By John H. Horwath, Ampco Pumps Co.

Page 231: Centrifugal Pumps Handbook

of the foot valve.After planning the required pip-

ing system, you should determinethe NPSH available of the system.(This is a desirable procedure to fol-low where your height plus dynamiclosses exceed 15 feet or handling ahot liquid or operation in a closedsystem.)

Keep in mind that a pump’s loca-tion can be of utmost importance inestablishing its suction requirements.Often you will find it more feasible torelocate the pump than to call for aspecial low NPSH pump. The cost ofthese changes must be weighedagainst the cost of a usually larger orspecial pump required for a loweravailable NPSH condition.

Net Positive Suction HeadNPSH (Net Positive Suction

Head) can be defined as the absolutepressure at a datum line (normallythe centerline of the impeller eye atthe suction nozzle) minus the vaporpressure of the liquid (at pumpingtemperature) being pumped.

Proposed SystemA. Proposed System

(units in feet of water) NSPH Available= hp ± (*) hz – hf – hvp where:hp = absolute pressure acting

on suction liquid surface hz = liquid suction height

above or below the pump impeller centerline

hf = total head losses in the suction including exit, entrance, fitting and pipe losses at the intended flow rate

hvp = vapor pressure of the pumped liquid at the pumping temperature

(*) for liquid level above center-line + applies;for liquid levelbelow centerline – applies

Existing SystemB. Existing System

(units in feet of water) NSPH Available = pg ± ps + v2s – hvp where:

2g

pg = gas pressure in closed

tank or atmospheric pres-sure in open tank

ps = gage pressure reading in the pump section center-line (steady flow condi-tions should exist at the gage tap; five to ten diam-eters of straight pipe of unvarying cross-section are necessary immediate-ly ahead of tap). The cor-rected gage reading is a minus term in the equa-tion if it is below atmos-pheric pressure.

v2s2g= velocity head at point

of measuring ps (based on actual internal diameter of pipe) at point of pressure taps

hvp= absolute vapor pressure of the pumped liquid at the pumping temperature

Pumped LiquidCharacteristics

The liquid being pumped canhave a profound effect on the pump’ssuction capability. Keep in mind thata centrifugal pump is not fully capa-ble of handling gases or vapors andthat it is always necessary for theabsolute pressure in any sector of thepump to be higher than the vaporpressure of the liquid being pumped.Entrained air also has an adverseeffect on pump performance. As littleas 1% entrained air by volume canreduce pump head and capacity sub-stantially. Under no circumstancesshould a standard centrifugal pumpunit be expected to handle more than3% air by volume as measured underpump suction conditions.

Substantial effort should bemade to keep air out of the liquidentering the pump. This commonlyoccurs if a vortex develops at the suc-tion pipe inlet and adequate submer-gence or effective baffling can helpprevent this condition. A leak in thesuction or pump stuffing box operat-ed under a vacuum may also intro-duce air into the pump’s suction. Awell-designed entry coupled withgood maintenance practices will alle-viate most entrained air situations.

Dissolved gas or gas evolvingfrom a chemical reaction can also betroublesome. Common practice callsfor the use of a large diameterimpeller (not necessarily a largerinlet) to meet the hydraulic perfor-mances based on handling cold clearwater alone. Where the percent of air(or gas) exceeds the recommendedlimit for standard pumps (which mayvary for different designs), it may beappropriate to consider less efficientpumps designed specifically for han-dling two-phase flow.

The boiling of a liquid that canoccur at reduced pressures is depen-dent on the liquid properties – pres-sure, temperature, latent heat ofvaporization and specific heat.

The stated suction characteris-tics of pumps based on cold water isstandard throughout the pumpindustry. Determination of other liq-uids must usually be made throughtesting, although the Hydraulic Insti-tute’s Standard 14th Edition (Figure70 in their publication) does provide

232 The Pump Handbook Series

Hp

Hf

+Hz

-Hz

Pg

Ps

Figure 2. Existing system design

Figure 1. Proposed system design

Page 232: Centrifugal Pumps Handbook

correction for some liquids at tem-peratures up to 400ºF. The sameoperating conditions as those withcold water are usually maintained inthe absence of data. Other contribut-ing factors are liquid surface tension,specific gravity and viscosity.

Some adjustment can be madefor certain liquids, however, to avoidany chance of cavitation. The sameoperating conditions as with waterare usually maintained. Alwaysremember that the pressure at anypoint within the pump must remainhigher than the vapor pressure of theliquid if cavitation is to be avoided.

CavitationWhen the absolute pressure

becomes equal to or less than thevapor pressure of the liquid beingpumped, bubbles consisting of dis-solved gases begin to form. The bub-bles are carried by the liquid flowinto an area beyond the leading edgesof the impeller vanes where higherpressure being developed causes thebubbles to condense and collapse,creating severe mechanical shocks.The term used to indicate thisprocess is “implosion.” The burstingbubbles begin to damage the pump’sinterior surfaces in the immediatevicinity, and this damage can bequite destructive over a period oftime, depending on the pressuresdeveloped, their collapsing rate andthe base material being subjected toattack.

Four common symptoms of cavi-tation are:

Noise – caused by the collapse ofvapor bubbles as they enter the highpressure area. This is typically iden-tified as a light hissing and crackingsound at the onset of cavitation and arotating noise when fully developed.

Vibration – caused by the impac-ting of the bursting bubbles on theimpeller surface, this condition canalso result in a premature bearingand shaft seal failure.

Drop in Efficiency – indicates theonset of a cavitating condition. Thedegree of efficiency drop-off increas-es precipitously as cavitation increas-es.

Erratic Flow – commonly occurs.The severity of the fluctuating flow isdetermined by the degree of cavita-

tion and pump design.

The PumpThe suction system must pro-

vide the energy to move the liquidinto the eye of the impeller. Thedetermination of a centrifugalpump’s NSPH requirement is estab-lished empirically through a series ofperformance tests run under specificconditions.

Of prime importance in the suc-tion sector is the pump’s inlet vaneangle. This is usually in the rangebetween 10 and 27 degrees indepen-dent of the pump’s design. An oftenused flow angle is 17 degrees – acompromise between efficiency anda low NSPH requirement. For bestefficiency the vane inlet angle wouldbe nearer 27 degrees while for a lowNSPH requirement one would gocloser to 10 degrees. Four to six vanesare commonly used in most designsof the smaller units available fromseveral pump vendors.

Manufacturers have developedvarious pump design innovations toimprove suction capability. Extreme-ly low NSPH availability may requirethe inlet blade angles to be reducedeven further. Some commercial unitsprovide a separate axial impeller(inducer) in the suction entry justahead of the main impeller to furtherinduce flow into the standardimpeller’s eye.

When a low angle inlet is intro-duced it may be necessary to reducethe number of vanes to decrease theblockage effect. Some manufacturersreduce the length of every secondinlet blade tip into the impeller pas-sage, speed (rpm) and profile cen-troid radius.

Pump manufacturers todayincorporate design features that pro-vide efficient conversion of velocityenergy to pressure energy, smoothcast surfaces for lower friction lossesand sharp vane tips to reduce entryshock losses in the impeller eye.Smooth surfaces can also defer theonset of cavitation because theydelay the formation of microscopicbubbles that form prematurely onrough surfaces prior to the inceptionof a cavitating condition.

Accurate determination of thestart of cavitation requires very care-

ful control of all factors that influ-ence pump operation. A number oftest points bracketing the point ofchange must be taken and the dataplotted. Because of the difficulty indetermining just when change starts,a drop in head of 3% is usuallyaccepted as evidence that cavitationis present.

SummaryAn important aid in diagnosing a

pump or application problem is amaintenance file card or folder kepton the pump that lists its history ofoperational problems as well as partsreplacement over the years. In caseswhere severe attacks occurred in theimpeller suction area caused by ero-sion or cavitation, photographs of theattacked area should be kept as wellas commentary describing operationat that point – be it noise, vibration,erratic flow and/or surges.

Become familiar with the termi-nology and procedures provided inthis article since they can aid in iden-tifying and resolving many of theproblems encountered on the suctionside of the pump. Understandingbasic pump suction concepts willenable you to more clearly explainyour situation to your pump repre-sentative when a persistent problemcan’t seem to be corrected.

When purchasing a new orreplacement pump for a sensitivesuction service, go beyond the price,pictorial and geographical presenta-tions and ask the following ques-tions:

1. Were the pump’s calibrationtests conducted in conformance withthe Hydraulic Institute’s standards?

2. Can the vendor provide per-formance data specific to your antici-pated suction requirements?

3. Can the vendor furnish a typi-cal pump impeller and casing set foryou to examine?■

John H. Horwath is a Senior Tech-nical Consultant with Ampco PumpsCompany in Milwaukee, WI.

The Pump Handbook Series 233

Page 233: Centrifugal Pumps Handbook

T here is a trend toward lowflow/high head pumps in thechemical processing industry.This article will focus on sev-

eral types of such pumps, describingthe advantages of each and tellingwhen they should be used. We willaddress the more popular lowflow/high head pumps, giving basicterms and discussing one specificdesign in greater detail.

Low Flow/High HeadMany industrial processes

require approximately 15 gpm with apressure of 200 to 1000+ ft TDH.Many companies use standard endsuction pumps in an attempt to meetthese needs. Most end suctionpumps, however, are not designedfor this type of application – a factthat leads to premature failure.

As an ANSI pump is operated tothe left of its Best Efficiency Point(BEP), several things happen. First,there is suction recirculation. Sec-ond, with even less flow, impellers,bearings and mechanical seals wearout faster. Last, in shutoff conditions,there is no flow and significant rise intemperature. All of these result inshorter pump life, less reliability,shorter seal life, and shorterMTBPM. However, several optionsare available to apply the right pumpto the right application.

Option 1 – Special ANSIDesign Impellers

Due to the large installed popu-lation of ANSI pumps in the world,many manufactures have designed

special type of impellers for low flowapplications. There are basically twoconfigurations.

The first is the “Barske” design.Basically, this is a semi-open impellerwith balance holes and it requires aspecial concentric casing. The advan-tage of this design is stable lowflow/high head applications, andonly the impeller and casing have tobe replaced on an existing ANSIpump for a conversion.

The second low flow impellerdesign is an enclosed type that pro-vides stable operation at low flowsbut has added benefits: the impelleris the only part that needs to bereplaced, and there are no wearrings. Thus, an existing ANSI pumpcan be changed out with minimalcost and time through a simplereplacement of the impeller. Figure 1illustrates this type of design.

Side Note: One company offers aheavy duty vertical enclosed-designchemical sump pump for lowflow/high head applications thatoffers an additional benefit. It has, asan option, balance holes in the backcover that allow the pump to runagainst a closed valve (dead headed)for extended periods of time with noadverse effect on pump life. This isideal for applications in which tanksare being used for a batch processthat only requires intermediateamounts of product. With thepump’s ability to run continuously,the customer can let the unit run andoperate a valve to meet require-ments. Also, the pump has ANSIimpellers and casings as a standard,which offers increased interchange-

ability of parts and less inventory inthe field.

Both basic impeller designs havebeen available for some time fromvarious pump companies. In addi-tion, these designs allow for fieldupgrades while keeping cost anddown time minimal. Furthermore, ifthese are purchased as new units, allother parts would be interchangeablewith other installed ANSI pumps.

These designs will handle asmall amount of entrained air and/orsolids that may be present. TheNPSHR is also less than standardANSI pumps running at the sameservice. These pumps are availablewith all the ANSI options –e.g., jack-eting and large bore stuffing boxesand a wide range of materials.

Option 2 - Pitot Tube Pumps

In a unique single stage centrifu-gal pump known as the Pitot tube

Pumping Options for LowFlow/High Head Applications

Familiarity with the major designs will help you match the right one to your application.

By Patrick Rienks, Sterling Fluid Systems (USA), Inc.

234 The Pump Handbook Series

CENTRIFUGAL PUMPS

HANDBOOK

Figure 1. Enclosed design for ANSIpumps

(cou

rtesy

of S

terlin

g Fl

uid

Syste

ms)

Page 234: Centrifugal Pumps Handbook

pump, the Pitot tube is stationary,and the inner casing rotates (See Fig-ure 2). The liquid enters the pumpthrough the suction line, passing themechanical seal, then enters therotor (Part 2) where it is brought upto rotor speed. The liquid near thelargest rotor diameter has a pressurethat obeys the basic mechanical lawsof centrifugal force.

A stationary wing-shaped Pitot-tube (Part 1), is placed inside therotor (Part 2) which has a circularopening near the largest rotor diame-ter. The Pitot-tube has a doublefunction. First, the liquid is forced toenter the tube due to lower pressurein the tube. Second, when the rotat-ing liquid hits the specially shapedstationary tube, the liquid speed istransformed into pressure energy.

This operational principleenables the pump to generate pulsa-tion-free flow and a stable NPSHRcurve.

The Pitot tube design offershigher pressures – up to 6200 feet ofTDH – whereas most ANSI pumpsare limited to about 900 feet of TDH.The pressure can be significantlyhigher in Pitot-tube pumps becausethe mechanical seal (Part 29) is sub-jected to suction pressure only, andthus the seal pressure is low, wherein an ANSI design the seal is subject-ed to a combination of suction anddischarge pressure.

Also, the Pitot tube design haslower NPSH requirements than theANSI style, and it can handle somesolids. Pitot-tube pumps offer greaterefficiency, 64% versus 35% for ANSIstyles. A wide range of alloys is avail-able, and the design provides stableflow in a wide range of applications.

And there is another advantage. Toincrease flow and head in most situa-tions, one needs to change only thePitot-tube and/or the speed. This pro-vides a quick and inexpensive solu-tion versus changing entire sizes ofANSI designs.

The Pitot tube design has beenin industry for more than 35 years.Thus, for special applications it is aproven problem solver. Typical appli-cations include cleaning, descaling,injection, boiler feed, process han-dling and hydraulic and spraying sys-tems in chemical, plastic, paper, steeland other industries.

Option 3 - Side Channel Pumps

The side channel design is usedmore in other parts of the world thanin the USA, but it is starting to drawdomestic interest. Photo 1 shows atypical side channel design. Figure 3shows a cross-sectional view.

This proven end suction multi-stage design brings the liquid into theeye of the impeller. The inlet stage iscentrifugal, and the impeller isenclosed, optimizing suction require-ments and providing superiorNPSHR values.

Additional performance isachieved by the open impeller stagesbehind the inlet stage, each of whichgenerates pressures many timesgreater than standard centrifugalpumps running at the same speed.The special configuration of thestages gives the pump excellent gas

handling capabilities and self prim-ing.

Also, unlike ANSI designs, sidechannel pumps will not vapor lock.ANSI will handle approximately 7%entrained gas vs. 50% for the sidechannel design. For example, thispump can handle volatile liquid up to150 gpm at 1200 feet TDH with anNPSHR of less than 1 foot. No otherdesign can accomplish this. Addition-ally, it can achieve suction lifts ofmore than 25 feet, has “floating”impellers that eliminate end thruston bearings, and can operate atlow speed and under 25” Hg vacuumor greater for long periods. Theside channel pump is a compactdesign available in a wide range ofmaterials.

Figure 4 illustrates the designprinciple of lateral channel stages inturbine pumps. This design can bemade in one to eight stages. From thediagram it can be seen that liquid or

vapor or mixtures enter the stagethrough the inlet port (A) in the suc-tion intermediate plate (B). Althoughnot shown, it should be noted thatthe internal face of this plate is flat,not channeled as in conventional tur-bine pumps. The mixture, once itencounters the rotating impeller (C),makes several regenerative passesthrough the unique side channeldesign shown (D), located in the dis-charge intermediate plate (E).

Due to centrifugal action, the liq-

The Pump Handbook Series 235

Photo 1. CEH side channel pump

(cou

rtesy

of S

terlin

g Fl

uid

Syste

ms)

Figure 3. Cross section of CEH sidechannel impeller stage

(cou

rtesy

of S

terlin

g Fl

uid

Syste

ms)

Figure 2. Combitube – Pitot-tube styledesign

(cou

rtesy

of S

terlin

g Fl

uid

Syste

ms)

4 17 31 2 1 26 29 36

Figure 4. Cross section of CEH sidechannel impeller stage

(courtesy of Sterling Fluid Systems)

E F G I H J

D C B A

Page 235: Centrifugal Pumps Handbook

uid, being the heavier component ina mixture, is forced toward theperiphery of the chamber, whereasthe lighter vapor or air collects nearthe center at the base of the impellerblades. Most of the liquid exitsthrough the discharge port (F), withthe remainder then guided along themini channel (G), which eventuallydead ends at point (H). There the liq-uid is forced to turn toward theimpeller hub, thus compressing anyvapor or entrained air at the base ofthe impeller blades. This compressedvapor or air is now forced throughthe secondary discharge point (I),after which it rejoins the major por-tion of the liquid that was dischargedthrough discharge port (F).

Thus, the problem of continualair or vapor build-up within thechamber has been overcome, and theliquid vapor mixture continues on tothe subsequent stage and is eventual-ly discharged from the pump.Where multiple stages are utilized,

they are staggered radially to bringabout balance and minimize shaftdeflection. The impeller in eachstage, although keyed for radial dri-ve, is allowed to “float” axially, thusassuming a running position in theequilibrium brought about by bal-ance holes (J), which are appropriate-ly positioned in the impeller hub.The floating action of the impelleralso eliminates axial thrust on thepump’s external ball bearings.

This design, also available in amag drive version, can be applied tothe same services as Pitot-tube andANSI style pumps, but it can be addi-tionally applied to self priming appli-cations. Another option is a barreldesign enclosing the entire pump foradded safety. However, due to thelateral gap between the vane wheelimpeller and the guide disk in thisdesign, these pumps should not beused where solids may be present.

The barrel pump is also availablein many materials, including non-

metallic PAEK.

SummaryThere are many possibilities for

low flow/high head applications. Wehave explained some basic guide-lines. However, if you have a specificapplication, its best to ask the appli-cations group of a pump company toreview the data in order to select theright pump for your needs. In today’schanging industry, there are manysolutions to help you keep yourprocess running, lower your costsand greatly increase your MTBPMby selecting the right pump for thejob.■

Patrick Rienks is a Product Man-ager for the Industrial Division of Ster-ling Fluid Systems (LaBour, Peerless,SIHI, and SPP Companies) in Indi-anapolis, IN. He earned his B.S. inChemical Engineering from Tri-StateUniversity and is currently enrolled atRose-Hulman Institute of Technologyfor his Master’s of Engineering.

236 The Pump Handbook Series

Page 236: Centrifugal Pumps Handbook

Introduction

Bearings used in centrifugalpumps are categorizedaccording to the direction ofthe forces they absorb, either

radial or axial thrust. Most small cen-trifugal pump bearings use anti-fric-tion bearings, ball or roller types,because they can be designed to han-dle a combination of both radial andaxial thrust loads.

Anti-friction bearings use ballsor rollers instead of a hydrodynamicfluid film to support a shaft load withminimal wear and friction (Figure 1).Cleanliness, accuracy and care arerequired when installing ball bear-ings. The ball bearing is a piece of

precision equip-ment manufactu-red to extremelyclose tolerances.To obtain themaximum servicefrom it, the shaftand housing mustalso be machinedto rigid toleran-ces. For example,locating shouldersmust be at rightangles to the shaftcenterline so thebearing will besquared with theshaft, and thehousing bores must be in almost per-fect alignment to ensure that thebearing will not be forced to operatein a twisted position. Maintenance ofball bearings is simple. Protect thebearing from contaminants andmoisture, and provide proper lubri-cation.

Bearing LoadsAssuming that the above condi-

tions are satisfied, the life of a ballbearing depends on the load it mustcarry and the speed of operation. Theloads on pump bearings are imposedby the radial and axial hydraulicforces acting on the impeller.

In any two-bearing system, oneof the bearings must be fixed axiallywhile the other is free to slide. Thisarrangement allows the shaft toexpand or contract without imposingaxial loads on the bearings. At thesame time, the arrangement locatesone end of the shaft relative to thestationary parts of the pump. Gener-ally, the outboard bearing of

between-bearings design pumps, orthe closest one to the coupling on aback pull-out design, is fixed axially.The inboard bearing is free to slidewithin the housing bore to accommo-date thermal expansion and contrac-tion of the shaft as shown in Figure2. Since the outboard bearing is fixedin the housing, it must carry bothaxial and radial thrust. The axialthrust is considered to be actingalong the centerline of the shaft andtherefore is the same at the outboardbearing as it is at the impeller. Theradial and axial loads combine to cre-ate an angular load at the outboardbearing.

Radial thrust acting on theimpeller creates radial loading onboth bearings. The magnitude of theload at each bearing can be deter-mined by the use of Figure 3 andthese equations:

P x aR1 = sR2 = P(a+s)

s

The Pump Handbook Series 237

CENTRIFUGAL PUMPS

HANDBOOK

Anti-Friction Bearings in Centrifugal Pumps

A detailed overview of bearing design, along with application and maintenance tips.By William E. (Ed) Nelson, Turbomachinery Consultant

1. Oil Formation

A–Bearing-static condition.

B–Partial operating speed.

C–Operating at full speed.

Figure 1. Theory of ball bearings

P

R2

R10.004"

0.002" 0.002" 0.002"

10"a 10"sImpeller

Mech.seal Bearing

radialBearing

axial

Figure 2. Loads on radial and thrust bearings

Page 237: Centrifugal Pumps Handbook

Where:R =Radial load on bearings 1 and

2 (pounds)P =Radial thrust on impeller

(pounds)a =Distance from the centerline

of the impeller to the centerline of the inboard radial bear-ing (inches)

s =Distance from the centerline of the inboard radial bearing to the centerline of the outboard bearing (inches)

Radial loads come from othersources as well. The weight of therotating assembly (shaft, sleeve andimpeller) gives one load. Unbalanceand external misalignment of theshaft give still another. The weight ofthe overhung coupling also creates abearing load. Pump designs shouldlimit the shaft deflections at the sealface to under 0.002 inch at the worstconditions. For single stage horizon-tal pumps, this will be with the max-imum impeller diameter at “shut off”conditions – i.e., closed discharge.For larger double suction pumps, thisload might well occur at the far endof the performance curve. Attentionpaid to cutwater clearances, Gap “B,”and “overfiling” of impeller vanescan reduce some of the hydraulicloads.

Ball Bearings TypesA ball bearing normally consists

of two hardened steel rings and sev-eral hardened balls utilizing a separa-tor to space the rolling elements andreduce friction as shown in Figure 3.The many ball bearing designs used

in industry are classified according tothe type of loading they receive: radi-al, thrust and combined. Sizes andclasses of precession of bearings aregoverned by the Anti-Friction Bear-ing Manufacturing Association (AFB-MA) and by the Annular BearingEngineers Committee (ABEC). Thereare five ABEC Classes, 1, 3, 5, 7 and9. Class 1 is standard, and Class 9 ishigh precision. Pump bearings gener-ally are Class 3, loose fit. ABEC Class9 is factory order only and has nolonger bearing life or higher speedrating than ABEC 1.

Three types of bearings are gen-erally used in centrifugal pumps:

Conrad Type – The Conrad type –identified also by its design featuresas the deep-groove or non-loadinggroove type – is the most widely used(Figure 4). A general purpose bear-ing, it is used on electric motors orwherever slight axial movement ofthe shaft is permissible. The deep-groove rings enable this bearing tocarry not only radial loading, forwhich it is primarily designed, butabout 75% of that amount of thrustload in either direction, in combina-tion with the radial load. API 610

Standard “Centrifu-gal Pumps for Petro-leum, Heavy DutyChemical and GasIndustry Services”pumps require thatsingle row or doublerow radial bearingsbe Conrad type withClass 3 or loose fit.This permits enoughflexibility to let theshaft correct for anymisalignment bet-ween the housingand the shaft.

Maximum Capacity Type– Anotherradial bearing is the maximumcapacity or filling-slot type. It is pro-vided with a filling notch thatextends through the ring or ringshoulders to the ring way, permittinga larger number of balls to be placedbetween the rings than can be donein the same size Conrad bearing. Thesupposed advantage of these bear-ings (larger load carrying capability)has vanished with the availability ofbetter steels and lubricants. Fillingslots often are not precisionmachined and can enter the ball con-tact area of the rings. This will resultin early bearing failure. Several bear-ing manufacturers consider a fillingslot bearing to be unreliable and dis-courage their use. The 1981 and latereditions of API 610 prohibit their usealthough they are still permitted byANSI specifications.

Angular Contact Type – Thisdesign allows the carrying of highradial loads in combination withthrust loads from 150 to 300 percentof the imposed radial load (Figure 5).Unlike the Conrad design, the con-tact angle is not perpendicular to thebearing axis. Several different anglesare available, offering a variety ofradial and thrust loadings. Fillingslots are not used in this design, and

238 The Pump Handbook Series

OUTER RING

SEAL ORSHIELD NOTCH

BOREOUTSIDEDIAMETER

INNERRING FACE

OUTER RINGBALL RACE

INNER RINGBALL RACE

OUTER RING FACE

WIDTH

CORNER RADIUS

SHOULDERS

CORNER RADIUS

SEPARATOR

Figure 3. Elements of a ball bearing

Figure 5. Angular contact bearing design

Figure 6. Back-to-back duplex mount-ing of angular contact bearings

Figure 4. Conrad bearing design

Page 238: Centrifugal Pumps Handbook

thrust loads can be imposed fromone direction only.

Duplex Type – These are identicalangular contact bearings placed sideby side. The contacting surfacesmust be ground to generate a speci-fied preload (Figures 6). This specialgrinding allows the two bearings toshare loads equally. Without it, onebearing in the pair would be over-loaded, the other underloaded.

API pump specifications requireduplex 40 degree contact angle thrustbearings mounted back- to-back witha light (100 pounds or 45 kg) preloadas the best choice. While the require-ment of a 7000 series, 40 degree con-tact angle, light preload bearingshould be a fairly tight specification,it is not. First, there are three 7000series bearing designs – light, medi-um and heavy. There is about a 50%change in capacity from one designto another. Second, some manufac-turers use more than one contactangle. The contact angle is the sourceof considerable confusion as shownin Figure 7. There are no standarddesignations to identify the 40 degreeangle.

Not all ball bear-ing companies manu-facture the 40 degreeangle angular contactbearing. If the localbearing supply houseis not on its toes,pump repairs couldbe made with non-specification bearings.The 40 degree contact

angle gives an 18 to 40 percentincrease in capacity over the 30degree angle depending on bearingsize. This differential is a very impor-tant piece of information, and unfor-tunately, it is communicated ordesignated in industry by a confu-sion of suffixes. The numerical codeused in bearing identification is most-ly standard among the various bear-ing manufacturers. However, thealphabetical prefixes and suffixes arenot. When identifying bearings fromcodes for the purpose of interchanging bearings, care should beexercised that the meaning of allnumbers and letters is determined soan exact substitution can be made.Most manufacturers supply cross-reference tables for identifying equi-valent bearings. A very good sourceis the bearing manual of the AFBMA.

MountingsThere are five types of duplex

angular contact bearing mountings,although only two are commonlyused. Rigidity of the shaft and bear-ing assembly depends in part on themoment arm between ball-contact

angles of duplex bearings. Withinreasonable limits, the longer themoment arm, the greater its resis-tance to misalignment.

API 610 calls for the DB or back-to-back mounting of angular contactbearings. They are placed so that thestamped backs of the outer rings aretogether. In this position the ball-con-tact angles diverge outwardly, awayfrom the bearing axis. With DB bear-ings the space between the divergingcontact angles is extended (Figure 8).Shaft rigidity and resistance to misalignment are correspondinglyincreased.

DF bearings are intended onlyfor face-to-face mounting. They areplaced so that the faces (or low shoul-ders) of the outer rings are together(Figure 9). Ball-contact angles thusconverge inwardly, toward the bear-ing axis. With DF bearings the spacebetween the converging contactangles is short. Bearing-shaft rigidityis relatively low. However, thisarrangement permits a greaterdegree of shaft misalignment thanother mounting methods. Some oldermultistage pumps use this mountingarrangement.

Special DesignsCentrifugal force in the un-

loaded thrust bearing causes its ballsto move out of their intended trackand operate on a skewed axis. Theballs begin to slide rather than rollduring rotation. The increased fric-tion that results reduces the viscosi-ty of the oil film, accelerates wear of

The Pump Handbook Series 239

Figure 7. Contact angles of various bearing manufacturers

Contact angles for 7000 series duplex bearingsContact angle

Bearingmanufacturer 20° 25° 30° 35° 40°

A X X XB XC X XD X

STAMPED FACESOF OUTER RINGS

TOGETHER

CLEARANCE BETWEEN

INNER RING FACES

Before Mounting Mounted

INNER RING FACESCLAMPED TOGETHER

THESE INNER ANDOUTER RING FACES

ARE FLUSH

Figure 8. Back-to-back (DB) mounting moment arm

UNSTAMPED FACESOF OUTER RINGS

TOGETHER

CLEARANCE BETWEENOUTER RING FACESTHESE INNER AND

OUTER RING FACESNOT FLUSH

Before MountingMounted

INNER AND OUTERRING FACES

CLAMPED TOGETHER

FACES FLUSHON BOTH SIDES

Figure 9. Face-to-face (DF) mounting moment arm

Page 239: Centrifugal Pumps Handbook

the raceways and leads to early fail-ure. Some recent work in bearingselection indicates that adherence tothe API specification of 40 degreebearings mounted DB may not beoptimal.

For single stage pumps in whichthe thrust action is steady and in onedirection at all flows, the use of a 40degree angular contact bearing toabsorb the primary thrust and a 15degree angular bearing for anyreverse thrust has extended the lifeof pump bearings. The 15 degreebearing decreases the tendency forball sliding and increased friction.The bearings also have machinedbronze retainers to reduce internalfriction further. The bearings arepackaged in pairs and are marked sothat when they are mounted theywill accommodate the primary thrustload (Figure 10). This arrangement isbetter in some but not all applica-tions.

Recommendations for bearinguse are summarized as follows:

A. Pump Type: Overhung singlesuction pump and low speeds, 1750cpm or below, with any load condi-tions

Bearing Style: 40 degree duplex, DB mounting, or 40 and 15 degree

B. Pump Type: Overhung singlesuction pump and high speeds, inexcess of 1750 cpm, with large thrustloads coupled with radial load

Bearing Style: 40 and 15 degree, DB mounting

C. Pump Type: Double suctionbetween bearings and high speeds, inexcess of 1750 cpm, mostly radialloads with low thrust loads

Bearing Style: Duplex 40 degree, DB mounting

Double-Row TypeEssentially, the double-row bear-

ing is an integral duplex pair of angu-lar contact bearings with built-inpreload (Figure 11). It resists radialloads, thrust loads or combined loadsfrom any direction. Two basic typesare available, corresponding to theface-to-face and the back-to-backmounting of conventional duplexbearings. Avoid using two double-row bearings on the same shaftbecause this makes the mounting toorigid. The bearings on each end ofthe shaft will tend to impose loads oneach other.

While double-row bearings canbe constructed with angular contact,such designs require a filling slot forassembly of at least one row. API 610prohibits the use of the designbecause of its greater vulnerability tofailure in reverse thrust applications.

Bearing MisalignmentCapability

The ability to tolerate misalign-ment between the bearing housing andthe shaft is dictated by ball and ringgeometry. Table 1 is a chart showingthe relative capabilities of three bear-ing types. Knowledge of these relativecapacities can help a maintenanceengineer or supervisor make substitu-tions to get out of many bearing prob-lems. Note the Conrad bearing’s ratedradial capacity and speed limit are taken as unity for comparison purposes.

The angular contact bearing cancarry almost double the Conrad’s radi-al rating in thrust. The Conrad canonly carry 75% of its radial rating inthrust, and the self-aligning ball bear-ing can carry only 20%. This meansthat thrust load on a self-aligning ball

bearing is prohibited. Note the angularmisalignment capability of the variousbearings. The Conrad can withstand15 minutes; the self-aligning ball cantake 16 times as much or 4 degrees.Values in this chart are for comparisonpurposes only. Actual catalog valuesfor load ratings and limiting speedshould be used.

Other BearingProblems

There are a number ofproblems associated withanti-friction bearings uti-lization that impact pumpreliability.

Figure 12. Load zones and retainers ofa ball bearing

Average relative ratingsType Radial Thrust Limiting Misalignment

speed

Conrad type 1.00 0.75 1.00 0 deg 15’Angular contact 40° 1.00 1.90 1.00 0 deg 2’Self-aligning 0.70 0.20 1.00 4 deg

+-+-+-

Table 1. Capacity for various bearing designs

Figure 10. Special design duplexangular contact bearings

Figure 11. Double row bearings

240 The Pump Handbook Series

Page 240: Centrifugal Pumps Handbook

RetainersA retainer ring or cage is used to

make all the balls of a bearing gothrough the load zone (Figure 12).The most common retainer materialis low carbon steel (1010 analysis)attached by fingers, rivets or spotwelding. Riveted or spot welded steelstrip retainers are more subject tofatigue failures. When a bearing ringor cage is misaligned, the balls aredriven up against the ring shoulder,the top ball to the left and the bottomball to the right. The center balls oneach side, at this particular point,tend to stay in the center of the ringbecause balls in this position relativeto the misalignment are not thrustloaded. The net effect of this action isto flex the ring in plane bending. Asthe inner ring turns, a cyclic retainerbending stress occurs.

The load on the retainer pocketis also cyclic. At the high thrust posi-tions, the retainer exerts the maxi-mum force in maintaining the ballspace. Since the retainer and ball arein rubbing contact, the thermal loadis at its highest on one side (no thrustload point). In this manner, theretainer is subjected to both a flexingand thermal cyclic load that can leadto fatigue cracking at retainer stress

points. Notches and rivet holes mayform. Due to the rubbing contactbetween the retainer and balls, thelubrication requirements here aremore critical than for the rolling con-tact between the balls and rings.Shock loading of the bearings alsocauses retainer failure at the pockets.

Some manufactures use pressedbrass, machined bronze, machinedphenolic and molded plastics in aneffort to reduce the heat generation.The relative desirability of bearingseparator types is as follows:

1. phenolic2. machined bronze3. pressed brass strips4. pressed steel strips5. riveted steel strips

Bearing CarriersIn order to use a larger radial

bearing and still be able to removethe impeller or mechanical seal fromthe shaft, some manufacturers utilizea bearing carrier similar to the oneshown in Figure 13 on between-bear-ing pumps.

The carrier is a shoulderedsleeve with a small clearancebetween it and the shaft. The radialbearing is then shrunk onto the out-side diameter of the carrier. Theproblem with this design is that if thebearing begins to heat up, due to lackof lubrication or some other reason,the carrier also heats up – expandinguntil it comes loose. At this point,even though the bearing has notfailed yet, the carrier may be free tospin on the shaft, a condition thatwill eventually cause the shaft to

bend or fail. Current API 610 specifi-cations require that the bearings bemounted directly on the shaft. Thereare a lot of carriers still in service.

Snap RingsSnap rings are flat, split washer-

like devices used by some manufac-turers to position componentsaxially – e.g.,ball bearings and sealsleeves on shafts.

As with bearing carriers, thereare two major problems in usingsnap rings. First, removal requiresthe use of a tool that is not normallyfound in the pump machinist’s toolbox. When used in a shaft, they mustbe positioned in a groove. The addi-tion of a radial groove in the shafteffectively reduces the diameter andmay weaken the shaft. When used ina bearing mounting, the rings permitconsiderable end float of the bearing.Current API 610 specification pro-hibits the use of snap ring mountedbearings.

Bearing ArrangementsDifferent arrangements of anti-

friction bearings can handle variousloadings imposed on the pump.Pump design is crucial in determin-ing possible bearing arrangements.

The Pump Handbook Series 241

Figure 14. Typical pump bearing arrangements – between bearing pump

StuffingBox

SuctionImpeller

Suction

Stuffing box

RadialBearing

Discharge

Thrustbearing

Figure 13. Typical bearing carrierdesign

Page 241: Centrifugal Pumps Handbook

Horizontal PumpsOverhung impeller pumps usu-

ally employ ball bearings only. In atypical bearing housing arrangement(Figure 2), the radial ball bearing islocated adjacent to the impeller orinboard position. It is arranged totake only radial loads. The thrustbearing is located closest to the cou-pling and usually consists of a duplexpair of angular contact bearings. Thebearings are mounted back-to-backso that axial thrust load can be car-ried in either direction. This duplexbearing pair carries both the unbal-anced axial thrust loading as well asradial load.

In between-bearing pumps, theball radial bearing and the ball thrustbearing combination have individualbearing housings (Figure 14). Theradial bearing is normally located atthe coupling end of the pump. Theball thrust bearing is located at theoutboard pump end. The thrust bear-ing must be secured axially on theshaft to transmit the axial thrust loadto the bearing housing through thebearing. The bearing is usually locat-ed against a shoulder on the shaftand locked in place by a bearing nut.This means that the shaft diameterunder the thrust bearing is less thanthe shaft diameter under the radialbearing. Thus, by mounting the radi-al bearing on the inboard (or cou-pling) end of the pump shaft, a largershaft diameter is available to trans-mit pump torque from the couplingto the impeller. The thrust bearing,on the other hand, is locked axially inthe thrust bearing housing, the radialbearing is axially loose in its housingto allow for axial thermal growth.

A popular combination forbetween-bearing double suctionpumps consists of journal type radialbearings and a ball thrust bearing. Insuch an arrangement, all radialpump loads are handled by the jour-nal radial bearing. The ball thrustbearing is mounted in the thrustbearing housing such that only axialloads are carried by the thrust bear-ing. The housing around the ballthrust bearing is radially loose. Ametallic strap is employed on theouter rings of the thrust bearing. Thisstrap locks into the bearing housing

to prevent rotation of the out-er rings. Such an bearingarrangement is useful in high-er horsepower and higherspeed applications where ballradial bearings would beimpractical due to speed, loadand lubrication limitations.Because the ball thrust bear-ing is located on the outboardend of the shaft, the shaftdiameter under the ball thrustbearing can be relativelysmall since no torque is trans-mitted from this end of theshaft.

Vertical PumpsMost vertical pumps dif-

fer from horizontal pumps inthat the entire axial thrust,consisting of axial hydraulicforces as well as the staticweight of the pump and thedriver rotor, is supported by the dri-ver thrust bearing. Therefore, the siz-ing of that bearing becomes a jointeffort between the pump manufac-turer, the driver manufacturer andthe end user.

Many end users require that thisbearing be rated to handle at leasttwice the maximum thrust load, upor down, developed by the pump in aworn condition with two times theinternal clearances it had when new.This requirement came into effectafter users experienced problems inthe field that result from the follow-ing facts or principles.

1. The calculation of pumpthrust is not highly accurate.

2. Pump thrust increases asinternal clearances increase.

3. The thrust load varies withthe vertical position of the impellerswith the casing(s).

4. The thrust load varies withflow. (In some cases it may evenreverse direction.)

A reasonable margin should beprovided between the driver thrustbearing rating and the maximum cal-culated pump thrust.

Motors for vertical pumps areavailable as solid shaft or hollowshaft units. On hollow shaft units,the pump shaft extends upwardthrough the motor shaft and is sup-

ported at the top of the motor shaft.Clearance is provided between theoutside diameter of the pump shaftand the bore of the motor shaft. Onsolid shaft units, a solid coupling isfurnished by the pump manufac-turer to provide rigid attachmentbetween the pump shaft and themotor shaft extension. Thus, thepump shaft is retained radially bythe lower motor bearing. This isgenerally considered to be a betterarrangement for most verticalpumps since the shaft runout willbe less and thus the seal or packinglife will be longer. Further, largerdiameter pump shafts can be cou-pled to solid shaft motors than canpass through the bore of hollowshaft motors, and this also providesincreased shaft rigidity.

Because of potential field prob-lems, the thrust bearing should bemounted in the driver top bearinghousing, farthest from the solidcoupling and the pump (Figure 15). In the event that a thrust bearingfails, any subsequent drop in thedriver/pump shafts could result ina mechanical seal failure that couldrelease hydrocarbons to the atmos-phere. Also, these could be ignitedby a hot bearing.

242 The Pump Handbook Series

Figure 15. Vertical pump motor – solid shaft

THRUSTBEARING

RADIALBEARING

Page 242: Centrifugal Pumps Handbook

Ball Bearing FitsUnfortunately, many pump

manufacturers do not indicate theproper bearing fits for shaft andhousings to guide shop repairs. Theoriginal dimensions of both the hous-ing and the shaft will change withtime due to oxidation, fretting, dam-age from locked bearings and othercauses. Every bearing handbook hastables to aid you in selecting fits. Thevibration effect of looseness on thebearing fits is different for the hous-ing and the shaft.

Housing Fits – Ball bearing fits inthe bearing housing are, of necessity,slightly loose for assembly. If thislooseness becomes excessive, vibra-tion at rotational speed and multiplefrequencies will result. Do not installbearings with O.D.’s outside of thegiven tolerance band since this mightresult in either excessive or inade-quate outer race looseness. Table 2shows rules of thumb.

Shaft Fits – A loose fit of the shaftto the bearing bore will give theeffect of an eccentric shaft, at a onetimes running frequency vibrationpattern. The objective of the shaft fitis to obtain a slight interference ofthe anti-friction bearing inner ringwhen mounted on the shaft. Thebearing bore should be measured toverify inner race bore dimensions.Do not install bearings with an I.D.

outside of the given tolerance bandsince this might result in eitherexcessive or inadequate shaft tight-ness. Table 3 gives rules of thumb.

Detection of Anti-FrictionBearing Defects

Anti-friction bearing defects aredifficult to detect in the early stagesof a failure because the resultingvibration is very low and the fre-quency is very high. If monitoring isperformed with simple instrumenta-tion, these low levels will not bedetected, and unexpected failureswill occur. The vibration frequenciestransmit well to the bearing housingbecause the bearings are stiff. Detec-tion of defects is best done usingaccelerometers or shock pulsemeters.

There are guidelines for evalua-tion of bearing deterioration. Forexample, a ball passes over defectson the inner race more often than

those on the outer becausethe linear distance aroundthe diameter is shorter.There are four dimensionsof a ball bearing that canbe used to establish somefeel for its condition:

1. A defect on outerrace (ball pass frequencyouter) occurs at about 40%of the number of ballstimes running speed.

2. A defect on innerrace (ball pass frequency inner) caus-es a frequency of about 60% of thenumber of balls multiplied by run-ning speed.

3. Ball defects (ball spin frequen-cy) are variable with lubrication,temperature and other factors.

4. Fundamental train frequency (retainer defect) occurs at lower thanrunning speed values.

A simple check for verificationof poor bearing condition ismade by shutting off thepump and observing thatthe high bearing frequencyremains as the pump speedreduces. This high frequen-cy signal will normallyremain until the pumpstops. The frequency indi-cation is normally from 5 to

50 times the running speed of themachine.

In next month’s conclusion tothis article we will examine variouslubrication methods for anti-frictionbearings, as well as the role of bear-ing protection devices, labyrinthsand magnetic seals.

PART IIThis second half of our series

examines ways to keep your anti-fric-tion bearings in top operating condition.Grease, oil flood, ring-oil and mist lubri-cation systems are all detailed, as wellas some advantages and disadvantagesof three bearing housing protectiondevices.Anti-friction pump bearings can beeither grease or oil lubricated. Fail-ure from lack of effective lubrication,either in type or quantity, constitutesa major source of bearing difficulties.The primary purpose of oil, or the oilconstituent of grease, is to establishan elastohydrodynamic film betweenthe bearing’s moving parts as shownin Figure 1. This film results from awedging action of the oil between theroller elements and raceways. Theformation of the film is, to a majordegree, a function of the bearingoperating speed and, to a lesserdegree, the magnitude of the appliedload. Lubrication for ball or rollertype bearings can be developed inthree ways:

1. full elastohydrodynamic oilfilm — no metal-to-metal contact

2. no separating oil film —metal-to-metal contact all the time

3. mixture of the above methods(boundary lubrication) — occasionalmetal-to-metal contact with an oilfilm present only part of the time

While the surfaces of bearingsare highly finished, there are, never-theless, small surface imperfections.Use of correct viscosity lubricantensures development of a full oil filmbetween rotating parts. In boundarylubrication, metal-to-metal contactoccurs, and friction wear develops. Ifthe bearings are operated with thecorrect viscosity lubricant for thespeeds and loads involved, a full elas-tohydrodynamic film will developbetween the rotating parts (Figure 1).

The Pump Handbook Series 243

1. Fit of bearing inner race bore to shaft is 0.0005 inch tight for small sizes: 0.00075 inch tight for large sizes.

2. Shaft shoulder tolerance for a thrust bearing is 0 to 0.0005 inch per inch of diameter off square up to a maximum of 0.001 inch.

Table 3. Rules of thumb for shaft fits

Table 2. Rules of thumb for housing fits

1. Bearing OD to housing clearance - About 0.00075 inch loose with 0.0015 inch maxi-mum.

2. Bearing housing out of round tolerance is 0.001 inch maximum.

3. Bearing housing shoulder tolerance for a thrust bearing is 0 to 0.0005 inch per inch of diameter off square up to a maximum of 0.002 inch.

Page 243: Centrifugal Pumps Handbook

Under these conditions the oil film isthick enough to separate the bearingscompletely, despite the unevennessof their surfaces.

Since there is no metal-to-metalcontact with full film lubrication,there is no wear on the bearing parts.The only time metal-to-metal contactoccurs is on startup, or when thebearing is brought to rest. A lubri-cant with a viscosity too low for theoperating loads and speeds permitsthe moving parts to penetrate the oilfilm, which results in their makingdirect contact. In boundary lubrica-tion, this metal-to-metal frictioncauses wear of the surfaces toincrease rapidly, as the film is fre-quently ruptured. Viscosity require-ments for both ball and roller typebearings are expressed in terms ofDN value, a factor used to comparethe speed effects of different sizedbearings. The DN value is obtainedby multiplying the bearing bore sizein millimeters by the actual rotatingspeeds in revolutions per minute.Once determined, the DN value ischecked against standard tables todetermine which viscosity oil to use.

It is important to remember,however, that the advantages ofproper viscosity can be offset by highspeeds if too much lubricant isplaced in the bearing and housingcavity. At high speeds, excessiveamounts of lubricant will churn theoil and increase the friction and oper-ating temperatures of the bearing.

Lubricants also help protecthighly finished bearing surfaces fromcorrosion and, in the case of grease,aid in the expulsion of foreign conta-minants from the bearing chamberthrough periodic regreasing.

Increased operating temperaturereduces oil viscosity and film thick-ness and accelerates deterioration ofthe lubricant. Petroleum based lubri-cants operated beyond the 200°Frange will experience a 50% reduc-tion in life for every 18 degrees abovethis level. At high temperatures, themore volatile components of oil andgrease begin to evaporate and cancarbonize or harden within the bear-ing cavity.

GreaseGrease lubrication is normally

limited to small, low horsepower,noncritical pumps which operate atrelatively low speeds and tempera-tures (Figure 16). The grease can belocated in the bearing housing sur-rounding the bearing or packed inthe bearing and then sealed.

Oil FloodedA more common lubrication sys-

tem for centrifugal pumps is the oilflood (Figure 17). In such an arrange-ment, the bearing housing provides a

sump, or oil reservoir. This sumpmaintains a level of oil at or near thecenterline of the lowest ball of thebearings, kept constant by means of aconstant level oiler. There are twoproblems associated with this type oflubrication. First, if the level is toohigh, frothing and foaming mayoccur, generating heat within thereservoir. Second, there is a verysmall range between the proper leveland the bottom of the balls, belowwhich the bearings are dry.

Ring-oiledRing-oiled systems are often

used to lubricate anti-friction bear-ings in larger horizontal pumpswhere, because of speed or loads, asimple flood system is not adequate.Rings of a diameter larger than theshaft ride on top of the shaft and dipinto the oil reservoir below (Figure18). These rings are located axiallybetween, but adjacent to, the bear-ings. The rotation of the shaft causesthe rings to rotate and carry oil fromthe reservoir up to the shaft. The oilis then thrown from the shaft by oilflingers, located next to each oil ring,directly into each bearing to assurecomplete lubrication. As the oil is cir-culated through the bearings, it isreturned to the oil sump. For propersystem function, the oil level in thereservoir must be maintained so thatat least 1/4 to 3/8 inch of the ringbore is immersed.

244 The Pump Handbook Series

Figure 16. Grease lubricated pumpbearings

CONSTANT LEVELOILER

LINE BEARINGHOUSING

OIL DRAINGROOVE

LINE BEARING

VENT

BEARING HOUSINGCOVER

SHAFT

WATERSHIELD

CASING

OIL DRAINPLUG

OIL LEVEL

Figure 17. Oil flood lubrication

Page 244: Centrifugal Pumps Handbook

The constant level oiler (Figure19) is a device used with both flood-ed and ring-oiled lubrication systemsand acts as a small reservoir for extra

oil while maintaining a predeter-mined level in the bearing housing.The constant level oiler uses a liquidseal to keep the oil in the bearingreservoir constant. When the oil lev-el in the bearing recedes due to con-sumption, the liquid seal on thespout is temporarily broken. This letsair from the air intake vents enter theoiler reservoir, which releases oiluntil the liquid seal and proper levelare re-established.

Unfortunately, each oiler instal-lation is slightly different, so somethought must go into setting thisposition. The proper level is usuallyindicated (by either a name plate,casting mark or stamp) on the side ofthe reservoir.

Oil Mist LubricationThe basic concept of the oil mist

lubrication system is dispersion of an

oil aerosol into the bearing housing.The equipment necessary for such asystem is shown in Figure 20. Thereare two types of inlet fittings orreclassifiers of the oil particles. Theydiffer in the degree of coalescencefrom essentially none for the puremist to partial for the purge mist asshown in Figures 21 and 22.

Pure MistIn pure mist lubrication (Figure

21), an oil/air mist is fed under pres-sure directly into the bearing hous-ing. There is no reservoir of oil in thehousing, and oil rings are notemployed; the pressurized mist flowsthrough the bearings. The movingcomponents of the ball bearings pro-duce internal turbulence, causingimpingement and collection of oilupon the surfaces of the ball bearing.Vents are located on the backside ofeach bearing to assure that the misttravels through the bearings, and adrain in the bottom of the bearinghousing prevents the buildup of con-densed oil as shown in Figure 23.

The advantage of pure mist isthat it creates an uncontaminatedenvironment in which the bearingscan operate, and protects them fromadverse environmental conditionswhile effectively eliminating heatbuildup. The oil mist system will fol-low the path of least resistance. Theback-to-back mounting of the angu-lar contact thrust bearing will havethe most windage, so most of theflow of a single inlet fitting will gothrough the radial bearing, whichhas less windage and hence lessresistance to flow. As a result, heavi-ly loaded bearings may require twospray inlets. The mist must flow fromthe inlet fitting through the bearing,then to the vent. For duplex angularcontact bearings, the flow should bein the same direction as the thrust(Figure 23). The position of the ventand the center spray can be inter-changed. Vent area should be abouttwice the reclassifier bore area. Thiswill create a slight back pressure inthe bearing housing, which keepsdirt out. All vents should carryapproximately an equal portion of airin multi-vent installations. Differentsized reclassifier orifices are needed

The Pump Handbook Series 245

CAP

SIGHT GLASS(FRONT)

STEM

MARKHERE

OIL RINGSHAFT

0.375" (9.5mm)

0.250" (6.3mm)

OIL LEVEL: • HIGH • LOW

Figure 18. Ring oiling

ADJUSTMENTRANGE

OUTERSLEEVE

LOCKSCREWS

OILLEVEL

Figure 19. Constant level oiler

Figure 20. Basic oil mist system

MISTHEADER

STD. GALV.PIPE

INSTRUMENTAIR SUPPLY

AIR FILTER

PUMP

BULK OILSUPPLY

RELIEFVALVE

OIL FILTER

RECLASSIFIER

MISTMONITOR

S.S. TUBE

S.S. TUBE-MALEPIPE CONNECTOR

REDUCER

SNAP DRAIN

VERTICAL PUMP

PUMP

MIST DISTRIBUTION MANIFOLD 2”GALV.

SNAPDRAINVALVE

TURBINEPUMP PUMP ELECTRIC

MOTOR

Page 245: Centrifugal Pumps Handbook

according to bearing size.During the mounting process,

bearings must be heated to about250°F to go on the interference fits ofthe shaft. Most of the oil on the bear-ing will flow off. To replace this oil,the pump bearing should be either:

1. Reoiled by filling the bearing

housing with oil up tothe shaft level. Theshaft should be turned 3or 4 revolutions so thatthe bearing is coatedand the oil drained outof the housing.

2. Connected to theoil mist system andoperated 8 to 12 hoursto “plate out” an oil filmon the bearing.

Purge MistAnother version of

oil mist lubrication iscalled purge mist (Fig-ure 22). This system isemployed in conjunc-tion with a conventionaloil flood or oil ringlubrication system. Itcombines the advan-tages of the positive oilcirculation created by

the oil rings or oil flood system withthe pressurized uncontaminated oilmist system. When this combinationis employed, a constant level oilerwith overflow feature is used to pre-vent buildup and flooding of thebearings, which can result in exces-sive heat buildup (Figure 24).

Bearing Housing ProtectionDevices

There is a close relationshipbetween the life of rolling elementbearings and mechanical seals inpumps. Liquid leakage from amechanical seal may cause the bear-ings to fail, while a rolling elementbearing in poor condition can reduce

seal life. Only about 10% of rollingelement bearings achieve their threeto five year design life. Rain, productleakage, debris and wash-downwater entering the bearing housingcontaminate the bearing lubricantand have a catastrophic effect onbearing life. A contamination level ofonly 0.002% water in the lubricating oil can reduce bearing life by asmuch as 48%. A level of 0.10% waterwill reduce bearing life by as muchas 90%. To improve the conditionsinside a bearing housing, varioustypes of end seals are used. In almostevery case, the normal operating lifeand quality of the end seal is notnearly as good as that of the rollingelement bearings. Improving thequality of the end seals will increasethe life of rolling element bearings.

Lip sealsThe advantages of the frequently

used lip seals are low initial cost,availability, and an easily understoodtechnology. New lip seals provideprotection in both static and dynamic modes. Their major disad-vantage is short protection life due towear of the elastomer. Life expectan-cy of a common single lip seal can beas low as 3,000 hours, or three tofour months. Thus, while a bearing isdesigned to last from three to fiveyears of continuous operation, the lipseal will provide protection for onlya few months. The temperature lim-its of lip seals are - 40°F to 400°F (-42°C to 203°C) for Viton.

LabyrinthsLabyrinths are devices that con-

246 The Pump Handbook Series

MIST INLET

PURE MIST

REMOVEOIL RING

VENT

OIL SIGHTBOTTLE REPLACE CONSTANT LEVEL

OILER WITH PLUGS

VENT

Figure 21. Oil mist lubrication - pure mist

MIST INLET

VENT

PURGE MISTFITTING

OIL SIGHTBOTTLE SEE CONSTANT LEVEL

OILER DETAIL

Figure 22. Oil mist lubrication - purge mist

SPRAY INLETS

VENT

DIRECTIONOF THRUST

Figure 23. Flow pattern of oil mist

CONSTANT LEVEL OILER

CL 3/10 DIA. HOLE

OPERATING OIL LEVEL

1/4 IN.

BEARINGHOUSING

Figure 24. Vented oil sight glass bottle

Page 246: Centrifugal Pumps Handbook

tain a tortuous path, making it diffi-cult for contaminants to enter thebearing housing. Unfortunately,there are both well designed andpoorly designed labyrinth seals. Theadvantages of labyrinths are theirnon-wearing and self-venting fea-tures. With no contacting parts towear out, a labyrinth can be reusedfor a number of equipment rebuilds.Because the labyrinth provides anopen, however difficult, path to theatmosphere, the bearing housingvent can be removed and the tappedhole plugged with a temperaturegauge.

The disadvantages of labyrinthsinclude a higher initial cost than lipseals and the existence of an openpath to the atmosphere, which canenable contamination of the lubri-cant by atmospheric condensate asthe housing chamber “breathes” dur-ing temperature fluctuations inhumid environments. Also, they donot work as well in a static mode asin a dynamic, rotating mode.

The temperature limits oflabyrinths are determined by theelastomers driving the rotor andholding the stator in place, the sameas for the lip seal.

Magnetic sealsMagnetic seals use a two-piece

end face mechanical seal with opti-cally flat seal faces held together bymagnetic attraction. They have a

design life equivalent to mechanicalseals and rolling element bearingsand can be repaired. The majoradvantage of magnetic seals is thehermetic seal they provide for thebearing housing. Because of the posi-tive seal, other arrangements mustbe made to allow for the “breathing”that results from expansion and con-traction of the air pocket above thelubricant during normal temperaturechanges. Disadvantages of magneticseals include higher initial cost and ashorter life than the almost infinitelife of a labyrinth. Magnetic seals aregenerally not recommended with drysump, oil mist lubricated bearinghousings or grease-lubricated bear-ings. The upper operating tempera-ture limit of magnetic seals is lowerthan that of labyrinth seals, in therange of 250°F (121°C).

Table 4 summarizes suggestionsfor the application of bearing hous-ing end sealing devices with varioustypes of lubrication systems andenvironments. Bearing life can beextended by improving the environ-ment of the bearing housing, and thiscan be accomplished simply byimproving the end seals.

ConclusionsThe effective working life of a

pumping system is influenced bymany factors that are not necessarilyapparent to the facility engineer.Look at all possibilities that can

cause a premature failure, not justthe bearings. However, anti-frictionbearings — their installation and theenvironment they operate in — are a major factor in pump life expect-ancy. Well cared for bearings canextend mean time between failures(MTBF).■

REFERENCES1. Dufour, J.W., Nelson, W.E., Cen-

trifugal Pump Sourcebook, McGraw-Hill, 1992.

2. SKF Staff, Bearings in CentrifugalPumps, An Application Handbook, SKFUSA Inc., 1993.

The Pump Handbook Series 247

Wet SumpPurge Oil MistPure Oil MistGrease

Vertical ShaftPositions

High Humidity orSteam withTemperatureCycling

Direct WaterImpingement

Dirt and DustAtmosphere

SUMMARY OF BEARING HOUSING PROTECTION DEVICESLip Seal Labyrinth Magnetic Seal

Possible Short LifePossible Short LifePossible Short Life

Acceptable

Acceptable

Possible Short Life

Possible Short Life

Very Short Life

AcceptableAcceptableAcceptableAcceptable

Top Position Only

Acceptable

Special Design

Acceptable

AcceptablePreferred

AcceptableNot Recommended

Top Position Only

Preferred

Preferred

Acceptable

Table 4. Summary of bearing housing protection devices

Page 247: Centrifugal Pumps Handbook

Often it happens that a userbuys a pump for a given sys-tem. Later its capacityproves to be inadequate, and

operations people request thatcapacity be doubled. A common mis-take in this situation is to purchaseanother pump of equal capacity toadd to the system while using thesame discharge piping configuration.After the new pump has beeninstalled, it becomes apparent toeveryone’s chagrin that the flow hasnot doubled.

The problem is that the operat-ing point has shifted because ofincreased friction losses in the dis-charge piping system due to higherfluid velocity.

The tools available to analyze apump system are the head-capacityor pump performance curves andthe system curve. Pump suppliersprovide head-capacity curves withtheir pumps as shown in Figure 1.On the y-axis they plot the head, orpressure. The x-axis shows the flow.Pump efficiency curves and oftenthe NPSH required are also shown.The buyer plots the system curve onan x-y axis as in Figure 2. This curverepresents the required dischargehead and the friction losses in thedischarge piping. The x-axis showsthe flow in gpm and the y-axis thefriction losses in feet. The desireddischarge head is a straight, horizon-tal line, since it remains constantfrom minimum to maximum flow.

When selecting a pump, thebuyer specifies the differential headand the capacity. This is the operat-

ing point of the pump. There aremany reasons to have pumps operat-ing in parallel. The most common isflexibility. If only one pump is usedfor a service and that pump breaksdown, the pumping system orprocess shuts down. A spare pump istherefore often mandatory in criticalsystems. The spare pump can beidentical to the main pump. Projectspecifications often demand 50%sparing; one or two pumps in paral-lel supply the total flow, with anadditional pump as a spare. Thisgives extra flexibility for mainte-nance and in case the process flowvaries.

Buying pumps for a new system

is fairly straightforward. The firststep is to determine the desired flow,the required head, the desired pipediameter and length, and the valvesand fittings included in the system.Next, plot a system curve. The y-axisshows head losses, the x-axis flow.Let’s say a particular applicationrequires 900 gpm at a constant headof 82 feet. Suction is flooded. Thisconstant head may be the pressureof a pipeline from which the liquid isto be pumped, or the elevation of areservoir, tank or pressure vessel.The optimum discharge pipe diame-ter is determined to be 8”. The laststep is to plot the friction lossesagainst flow.

248 The Pump Handbook Series

CENTRIFUGAL PUMPS

HANDBOOK

Centrifugal PumpsOperating in Parallel

When it comes to pumps and flow, one plus one doesn’t necessarily equal two.

by Uno Wahren, Consultant

100

170

160

150

140

130

120

110

100

90

80

70

60

50

40

30

20

10

HEAD IN

FT

14

12

8

6

4

2

0

NPSH

60

50

40

30

NPSH

FLOW IN 100 USGPM

1 2 3 4 5 6 7 8 9 10 12 14 16 18 20 22 24 26

HEAD CAPACITY CURVE21" DIA

66%74%

80%

19" DIA.

17" DIA

OPERATING POINT

84%

86%

84%

80%

74%

66%

Figure 1. Typical head-capacity curve

Page 248: Centrifugal Pumps Handbook

Figure 3 shows a straight lineparallel to the y-axis drawn at 900gpm and intersecting the systemcurve at 82 ft. This particular systemrequires a pump that delivers 900gpm with a discharge head of 82 ft.Two pumps each delivering 450gpm, or three pumps each delivering300 gpm, will achieve the sameresult. In this case, the decision is touse two pumps running in parallel,each with a capacity of 450 gpm.

Each pump operating alonedelivers approximately 520 gpm atonly 68 ft. This is adequate becausethe end of the line or static head is60 ft. If the calculated system headcurve is above actual, it will be nec-essary to throttle the pump flow,using either the discharge isolationvalve or an eventual control valve onthe discharge line.

Suppose a system has one pumpdelivering 450 gpm against 68 ft. Thepump discharge line is 6” new steelpipe. The fluid is water with a spe-cific gravity of 1.0. There is a requestto double the flow. In a case like thisit is very common to request anotheridentical pump — one that will deliv-er 450 gpm at 68 ft. That is a mis-take. With the addition of the second450 gpm pump, the combined (1 and2) head-capacity curve will bisect thesystem curve (Point B) at 790 gpm.The total head required at that flowis 78 ft as shown in Figure 4.

If the new pump is alreadybought and installed, flow can beincreased by changing the dischargepiping from 6” to 8”. This is usuallynot an economical solution. The dis-charge line may be long. Replacingvalves and fittings may be expensive.Another solution is to buy a thirdpump and run the three pumps inparallel. This can also be expensiveand adds to maintenance costs.Changing the impellers to a largerdiameter that will still fit into thepump casing may also solve theproblem.

The characteristics of a centrifu-gal pump are such that with constantspeed (rpm) and a specific impellerdiameter, the curve will not change,regardless of the properties of theliquid being pumped. However, thecurve will change if either the speedor the impeller’s diameter change.

The performance curve shownin Figure 1 will change if the pump’srpm changes from 3560 to 4200through a speed increasing gear asfollows:

Q1 = 450 gpm; H1 = 160 ftQ2 = 450 x 4200/3560 = 530 gpm H2 = 160 x (4200/3560)2 = 223 ft

Increasing pump speed isuncommon. Pump characteristicschange when speeds are higher thandesign specifications. As NPSHrequirements increase, internal

recirculation can become a problem.The shaft and bearings may not bedesigned for the increased torqueand loads of the higher speed andhorsepower requirements. A geartrain means added maintenanceproblems. A larger driver may berequired. Why get into a situationlike that?

Assuming that the installedimpellers have an outside diameterof 6”, the maximum size impeller forthat pump is 6.3” O.D. The affinitylaw for a constant speed pump is:

D1/D2 = Q1/Q2 = √H1/√H2

The flow and head for a 6.3 diam-ter impeller are:

Q2 = (450 x 6.3)/6 = 472 gpmH2 = (√68 x 6.3/6)2 = 75 ft

(Note: Some impeller designs donot precisely agree with the affinitylaws for impeller diameter changes.Always discuss this with the pumpmanufacturer.)

As shown in Figure 5, the twopumps with the larger impellersoperating in parallel still do notdeliver the full 900 gpm. In this case,the total flow might be enough forthe particular process. On the otherhand, it may not. Increasing theimpeller size will often not solve theproblem.

To buy a pump that will doublethe flow requires plotting the systemcurve against the head-capacitycurves. The plot shows that at theincreased flow the system willrequire a total discharge head of 82ft. When the first pump (pump 1)runs alone on the system curve, theflow is 450 gpm and the head is 68 ft.Next select a pump (pump 2) which,running in parallel with pump 1,will deliver the required flow at therequired total discharge head (TDH).For this example, an adequate pumpis one with a flow of 645 gpm with aTDH of 72 ft. The pumps are operat-ing in parallel on the same systemcurve. The pumps are sized so thatpump 1’s head-capacity curve inter-sects the system curve before pointA. Therefore, pump 2 is the com-manding pump. If pump 1 starts first(before pump 2) it will back off andthe system will deliver only about650 gpm against 72 ft of head.

Figure 6 shows how the two dis-

The Pump Handbook Series 249

170

150

130

110

90

70

50

30

10

HEAD INFT

FLOW IN 100 USGPM

1 2 3 4 5 6 7 8 9 10 11 12

SYSTEM CURVE

STATIC HEAD

Figure 2. System curve

Page 249: Centrifugal Pumps Handbook

similar pumps can operate correctlyonly if the head-capacity curve inter-sects the system curve on the ADportion of the head-capacity curve.The pumps may be throttled, but notfurther back on their curve than 525gpm (Point A), for then pump 1 willback off, letting pump 2 deliver onlyits full capacity. For the system tooperate, pump 2 has to start first.After pump 2 has reached full flow,pump 1 can be started. (Figure 6).

When selecting an additionalpump, bear in mind that no pumpwill operate satisfactorily from zeroflow to the end of curve flow. Pumpmanufacturers show minimumallowable constant flow on theirpump curves. This means that belowthat flow, the pump is unstable. The

minimum flow requirements for lowhead, low capacity pumps rangefrom 20 to 25% of flow at Best Effi-ciency Point (BEP). It is not uncom-mon for large multi-staged pumps tohave minimum flow requirements ashigh as 65% of BEP capacity.

Always pick a pump where therequired flow is less than the flow atBEP. It is good practice to specifythat a pump will operate satisfactori-ly at 120% of the rated flow, sinceflow requirement may increase. Thehigher the flow, the higher the NPSHrequirement. At some point beyondthe BEP, the pump curve will col-lapse. Operations beyond that pointwill cause cavitation and severepump damage.■

Uno Wahren is a Registered Pro-fessional Engineer in the State of Texaswith over 30 years of experience as aproject and rotating equipment engi-neer in the oil and gas industry. He hasa B.S. degree in Industrial Engineeringfrom the University of Houston, and aB.A. in Foreign Trade from the Thun-derbird Graduate School of Interna-tional Management, Glendale, AZ.

250 The Pump Handbook Series

170

150

130

110

90

70

50

30

10

HEAD INFT

FLOW IN 100 USGPM

1 2 3 4 5 6 7 8 9 10 11 12

STATIC HEAD

SYSTEM CURVE

CURVE 1 + CURVE 2

TWO PUMPS

CURVE 1

PUMP 1 ONLY

Figure 3. Two pumps in parallel

170

150

130

110

90

70

50

30

10

HEAD INFT

FLOW IN 100 USGPM

1 2 3 4 5 6 7 8 9 10 11 12

STATIC HEAD

SYSTEM CURVECURVE 1 + CURVE 2

TWO PUMPS

CURVE 1

ONE PUMP

A

B

Figure 4. Addition of identical pump

170

150

130

110

90

70

50

30

10

HEAD INFT

FLOW IN 100 USGPM

1 2 3 4 5 6 7 8 9 10 11 12

STATIC HEAD

SYSTEM CURVE

CURVE 1 + CURVE 2

TWO PUMPS

CURVE 1

ONE PUMP

Figure 5. Larger diameter impeller pumps operating in parallel

170

150

130

110

90

70

50

30

10

HEAD INFT

FLOW IN 100 USGPM

1 2 3 4 5 6 7 8 9 10 11 12

STATIC HEAD

SYSTEM CURVE

PUMP 2

PUMP 1 A

D

C

B

Figure 6. Two dissimilar pumps operating in parallel

Page 250: Centrifugal Pumps Handbook

Point of view is everythingwhen discussing fire pumpsystems. For the engineering-contractor, they are relatively

uncommon systems referencingspecifications unheard of in conven-tional process units. To the purchas-er, they are mandated systems thattake time, space and capital awayfrom money-making units. Forusers, fire pump systems are (orshould be) a once a week testrequirement.

But in spite of the time, space andmoney constraints, fire pump sys-tems must work as required. Period.Hundreds of lives and millions ofdollars in hardware and productioncosts rely on the performance ofthese systems. Fortunately, theyalmost always work.

Because of the critical nature ofthis service, one might think thatindustry standards could simply beinvoked to insure reliable designand specification. Of course theycan. They just never are. At least notwithout the addition of supplemen-tal proprietary specifications thatcan conflict with industry standards,government regulations and some-times with the basic system design.Seemingly minor requirements canresult in the loss of a listing agencylabel and render a perfectly func-tional system unacceptable to insur-ers or governmental agencies.

PumpsThe basic fire pump system

includes a UL (Underwriters

Laboratory) or FM (Factory Mutual)listed pump, driver and controller.Pumps are rated in discrete incre-ments starting at 25 gpm and extend-ing through 5000 gpm. Each capacitydesignation is tested for compliancewith NFPA 20 (National FireProtection Association) requirementsfor performance and by independenttesting institutions such as UL or FMfor design, reliability and safety. Asingle pump can be certified formore than one operating point, but itmust meet all performance anddesign criteria at both points.

In the area of basic hydraulic per-formance, a pump must deliver atleast 65% of rated discharge pres-sure at 150% of rated flow toachieve NFPA 20 acceptance. So a1500 gpm pump rated at 150 psigmust deliver 2250 gpm at a dis-charge pressure of not less than 97.5psig. This operating range thenestablishes the design parametersfor the piping system and fire fight-ing equipment. Another require-ment establishes that the maximumshut-off head must not exceed 140%of design head.

Both horizontal and vertical cen-trifugal pumps are available as list-ed pumps. (For brevity, we will usethe term "listed" to refer to anyequipment certified by UL, FM, theCanadian Standards Association(CSA) or other agency as suitable forfire pump system application.)While there are some obviousdesign differences between therequirements for these two styles,

the performance prerequisitesremain un-changed. It is importantto note that listed pumps are notdesigned or manufactured to API610 standards.

Materials of construction varywith the type of system and the firefighting fluid. The standard cast ironcase with bronze impellers and wearsurfaces is the most common in land-based applications. This material isgenerally suitable for fresh water andsometimes brackish or even saltwater services, depending on thelength of service and anticipated lifeof the unit. Even in fresh water ser-vices, it is critical to consider casingcorrosion rates when choosing mate-rials. For more aggressive fluids,many manufacturers offer variousgrades of bronze, including zinc-freeand nickel-aluminum bronze. Higheralloys of 300 series stainless steel andAlloy 20 may also be available forvery corrosive services or situationsin which the expected project life isextremely long.

The Pump Handbook Series 251

Fire Pump Systems-Design and Specification

Think your fire pumps are just like the rest of your fluid-movers? Think again.Rigorous standards and certifications make sure these life-savers are up to snuff.

by George Lingenfelder and Paul Shank, Precision Powered Products

CENTRIFUGAL PUMPS

HANDBOOK

Photo 1. Fire water pump with dieselengine and air start with a nitrogenback-up system

Page 251: Centrifugal Pumps Handbook

Vertical pumps are available in amuch wider range of metallurgiesbecause of their applications in cor-rosive offshore environments.Various grades of bronze are com-mon, as are higher alloys. Carbonsteel is very rare. Material optionsfor horizontal units include bronzecasing materials, moving up to car-bon steel and then austenitic steelsfor both case and internals. As withvertical pumps, higher alloys arealso available for special situations.

From a commercial standpoint,cast iron-bronze fitted (CIBF) pumpsare the basic choice. Standard alu-minum bronze materials such asASTM B148 Alloy 952 (for verticalunits) generally command price pre-miums in the 4.5-5 multiplier range.Nickel-aluminum bronze (B148 Alloy955) will increase this pricing by anadditional 10-15%. Carbon steel (forhorizontal units) will raise standardCIBF costs by 3-4 times. Stainlesssteel, typically 316SS, increases basecosts by a factor of 5-8 dependingon size. Higher alloys can increaseCIBF costs by up to 8 times.

Auxiliary equipment and acces-sories offer many possible optionsfor the pump. While these will bediscussed in greater detail later, it isimportant to note that most firewater pump manufacturers only

offer a limited num-ber of materials andoptions for listedpumps. Pumps (likedrivers and otherequipment) can belisted for fire pumpservice only whenthey are manufac-tured in the samematerials and config-

uration as the design tested by thecertifying agency. Changes in casingmaterials, even when poured by thesame foundry using the same pat-terns, can prevent a manufacturerfrom labeling the pump. Likewise,changing accessories, especially onengines, can result in the loss of thelabel.

In practice, this is usually asemantic issue and rarely becomes asignificant problem, provided thatthe original design or unit waslabeled and that the changes areclearly upgrades in materials, con-trol and/or reliability. The impor-tance of labeling also varies with the location and type of fire pumpservice, as well as responsible gov-ernment agencies and insurance car-riers. Most manufacturers offerlabeled equipment in the most ele-mental fashion for application in thewidest range of markets. This makesit possible for the specifying engi-neer and user to work together withthe supplier to develop the best sys-tem for their application. Even

though the end user might cus-tomize the system and add acces-sories that void the label, the origi-nal equipment was labeled, and thisshows the intent of the user toinstall labeled equipment.

DriversWhile the pump is the mechani-

cal heart of a fire water system, thedriver is a critical component, fre-quently requiring more attention inspecification and design. Dieselengines power the vast majority offire water pumps, creating anotherdifficulty for the specifying engineer.Most of us have limited experiencewith this category of equipment.They are rare in general plant designand frequently used only becausethe power required exceeds theavailable UPS (UninterruptiblePower Source). Unlike the situationwith motors, the specification andselection of diesel engines cannot befoisted off on another discipline.

The basic, listed fire pumpengine, offered by many manufac-turers, is more than adequate interms of reliability and service life.After all, this system will only oper-ate once a week for thirty minutesuntil it is called on to run in a realemergency. Then it will operate foreight hours (the standard fuel tanksizing) or until it is destroyed in theconflagration.

A great deal of time and energy isexpended on the starting system,

252 The Pump Handbook Series

Figure 1. NFPA 20 vertical fire pump sys-tem with right angle gear, diesel enginedrive and elevated fuel tank with contain-ment reservoir

Figure 2. Plan view of horizontal fire water pump system with dieselengine drive, discharge piping main relief valve and discharge cone

Page 252: Centrifugal Pumps Handbook

which often includes back-up startsystems and sometimes even multi-ple starting systems. Basic systemsinclude dual sets of batteries. Alsocommonly available are pneumaticsystems—compressed air with bot-tled nitrogen back-up and hydraulicstarting systems. Many of the largerengines, above 200 hp, offer twoports for starters, so that two typesof starting systems can be used.This having been said, most experi-enced engine users agree that if theunit does not start within the firstthree to five seconds, the likelihoodof starting at all is just about nil. Forthis reason, it is very important topay close attention to the startingcycle of the unit during the weeklyexercise sessions. Remedy any diffi-culties or malfunctions at once.

Although relatively uncommon,some systems also use motors, espe-

cially when the firewater pumps arealso used for water lift or wash-down services. Whether they aremore reliable remains an open dis-cussion. The simple fact of theirpredominance as drivers has madethem more acceptable. Like othermajor components of fire pump sys-tems, motors are available inUL/FM and CSA listed varietiesfrom a wide range of manufactur-ers. While still not widely accepted,the most recent revision of NFPA 20(January 1998) now requires firepump motors to be UL listed.

ControllersControllers were the last major

component of fire pump systems toreceive certification by various list-ing agencies. Unlike other compo-nents, however, they come with awide range of options and even pro-vide remote system contacts for theother pump, driver and otheroptions that may need to be incor-porated into the control scheme.

A few areas in the design andspecification of the controllershould be reviewed carefully. Thefirst is area classification. Unusualas it may seem to designers andspecifiers who are unaccustomed toit, standard drivers and controllersare not rated for any NationalElectrical Code (NEC) area classifi-cation. Depending on the manufac-turer, they can add this as an optionor build a custom controller thatmeets a specific area’s classification.Meeting a Division I or II classifica-tion usually involves adding a Z-purge system, which can be expen-sive relative to the cost of the con-troller (typically increasing cost byup to 50%), but this is currently themost cost-effective method to meetthese requirements. Custom con-trollers, a very expensive alterna-tive, will generally incorporate her-metically sealed or NEMA 7 enclo-sures for all potentially arcingdevices. In addition, diesel enginecontrol sensors and battery startsystems are not intended for a clas-sified area. Electric motors mustalso be specified for the area.

Pneumatic engine controllers with

air start systems are an obviousalternative when area classificationbecomes an issue. Although not com-mercially available as UL/FM listedunits, it is generally agreed that theyare inherently explosion-proofbecause they have no electrical com-ponents. Custom built electric con-trollers, however, can meet the areaclassification requirements of anyfacility. Controllers incorporatingPLC logic are also available as stan-dard commercially available, but notas UL/FM listed units, or pneumaticcontrols as custom built units.

Custom manufactured controllersare becoming more common, espe-cially in the HPI/CPI markets tomeet NEC area classifications. It isimportant to note, however, thatwhile these units may consistentirely of UL (or other) listed com-ponents, this does not convey a ULlisting to the controller. Thus, thecontroller does not meet the UL218requirements. This may appear tobe a subtle differentiation, but localcodes and regulations, as well asinsurance requirements, may req-uire sacrificing an area classificationfor a UL/FM listing.

InstrumentationFire pump systems are not

process systems. Thus, much of theadvanced instrumentation applied toprocess systems is relatively uncom-mon in the world of fire pumping.However, some systems includetransmitters and other “smart”devices that provide additional infor-mation to the user. Carefully reviewthe purpose and function of thesedevices to determine their valueunder the anticipated operating con-ditions and to insure that they pro-vide valuable information withoutcontributing to needless complica-tion. Unlike a process system,where almost every eventuality canbe understood, evaluated and pre-pared for, the full array of condi-tions under which a fire pump sys-tem will be used is almost impossi-ble to imagine. Controls should beas simple and straightforward aspossible, lest they fall victim to thelaw of unintended consequences.

The Pump Handbook Series 253

Photo 3. Vertical fire water pump foroffshore installation. Compressed airstart with back-up nitrogen systemand expansion receiver

Photo 2. Custom fire water pump sys-tem designed for Class 1, Groups C &D and Division II. The controller fea-tures all NEMA 7 switches mounted ina stainless steel cabinet. The batterycharger and battery case are purgedand constructed of 316SS. The sys-tem has a primary electric start with aback-up hydraulic system.

Page 253: Centrifugal Pumps Handbook

Industry StandardsTwo major organizations are the

primary promulgators of standardsfor fire pump systems, and theirroles, far from being conflicting, arecomplementary.

National Fire ProtectionAssociation

The National Fire ProtectionAssociation, through its NFPA 20Standard for the Installation ofCentrifugal Fire Pumps, providesthe most complete set of provisionsfor the design and installation ofvarious components as well as theoverall system. The NFPA 20 setsstandards for design and construc-tion of all the major components inthe fire pump system, includingminimum pipe sizing tables, elec-tric motor characteristics, perfor-mance testing, and periodic testingand system design. NFPA 20 fur-ther attempts to develop a safetystandard or level of performancefor centrifugal fire pump systemsto provide a reasonable degree ofprotection for life and property.Under these provisions, alternatearrangements and new technolo-gies are permitted—and in factencouraged.

The National Fire ProtectionAssociation does not, however, certi-fy or evaluate compliance with itsvarious specifications and guide-lines.

NFPA 20, while the source ofsome performance requirements, isbest known for its detailed designprovisions for all components com-monly found in fire pump systems.

Unlike many industry standardsthat address only the design, manu-facture and testing of specific com-ponents, NFPA 20 is a veritable'how-to' manual for the engineer oruser who needs to develop guide-lines and specifications for firewater systems. Taken as a completedocument with referenced texts,this standard alone will assure thepurchase and installation of a reli-able fire pump system. Take care indeveloping additional specifications,whether stand-alone or ancillary, toavoid conflicts with NFPA 20.Because of the detailed nature ofthis specification, add-on require-ments can frequently have the effectof actually diminishing the stan-dards.

Here again it is important to keepin mind the function of the equip-ment in a fire pump system.Contrary to the design considera-tions in process units, fire pumpsystems should be designed to oper-ate reliably under the most adverseconditions for as long as required,but typically this is only a matter ofhours. In a process unit, systems aredesigned to protect themselves withvarious shutdowns and monitors.Fire pump systems are designed toprotect the plant and personneleven if that means operating todestruction. In other words, threehours into a fire fighting event, thevibration level of the diesel enginedriver is of no significance so long asthe pump continues to deliver ade-quate flow.

NFPA 20 includes pump designguidelines for vertical shaft turbinepumps, horizontal (both end-suctionand split-case) and vertical in-linepumps. It also includes standardsfor motors, both horizontal and ver-tical, right angle gears and dieselengine drives.

Additionally, the guidelines arean excellent resource for auxiliaryand ancillary equipment found inmost fire water pump systems. Forexample, the diesel engine specifica-tions include not only requirementsfor the engine itself, but also thefuel supply, exhaust system andcontrol system operation.

Underwriters LaboratoriesUnderwriters Laboratories (UL)

reviews equipment and systemsfrom a performance orientation.The organization provides an inde-pendent, third party evaluation ofmanufacturers for a variety of com-ponents, but for fire pump systemsthe primary specifications for ourconsideration are UL448 for pumps,UL1247 for diesel engines andUL218 for controllers.

In certifying equipment to thesestandards, UL reviews constructiondesign, materials of constructionand overall performance. Thesereviews assess the ability of theequipment to perform the task forwhich it is to be rated. Followinginitial certification, UL maintains anongoing surveillance program toinsure continued adherence todesign and performance criteria. ULfield representatives make unan-nounced visits to all manufacturersdisplaying the UL label. If teammembers encounter problems orquestions, they may schedule morefrequent visits.

UL standards are among the moststringent in the industry, and equip-ment certification is necessarilyvery specific. So what happenswhen a listed product is modified tomeet customer specifications or sitespecific requirements? If the changeis clearly an upgrade to the existinglisted design and not a majorchange in the product, UL will fre-quently certify the modificationwithout re-testing. Major modifica-tions, of course, must go throughthe complete certification process.The manufacturer is responsible forinitiating changes to the listing.

UL448 sets the standards for cer-tification of centrifugal pumps foruse in fire water systems. As previ-ously noted, UL448 is based on theability of the unit to perform underthe conditions of service requiredand according to the manufacturer'sspecifications with regard to totaldifferential head (TDH), capacity,and efficiency or power require-ment. Materials of construction arereviewed for strength and corrosionresistance. Each unit or size to be

254 The Pump Handbook Series

Photo 4. A 570 hp V-12 diesel engine driverfor all nickel aluminum bronze fire waterpump producing 3500 gpm at 180 psi.

Page 254: Centrifugal Pumps Handbook

certified is given an operational per-formance test and is also hydrotest-ed to twice the manufacturer's pub-lished Maximum AllowableWorking Pressure. This is a substan-tial increase over pump industry stan-dards of 150% hydrotest pressures.

After certification, UL448 req-uires that each unit bearing the ULmark be tested successfully forhydrostatic integrity and hydraulicperformance.

Diesel engines fall under thescrutiny of UL1247. Again, theemphasis is on performance.Certification requirements includeextended testing on dynamometersas well as speed control. Because ofthe importance of driver speed incentrifugal pump applications,speed control and overspeed shut-down operation are critical areas.Units are also tested for their abilityto start under a wide range of condi-tions, both hot and cold.

After certification, each produc-tion unit shipped must be subjectedto a dynamometer test includingperformance checks of the speedcontrol and overspeed shutdownsystems.

Engine and motor controller spec-ifications are covered by UL218,which is written and administeredby the Industrial Controls section ofUnderwriters Laboratories. Whilethis section is closely aligned withUL's primary purpose to reviewequipment for fire and shock haz-ards, the critical nature of fire pumpcontrollers makes performance animportant concern of this standard.These units are reviewed for safety,of course, but the importance ofstarting a fire water pump underemergency conditions warrants adifferent philosophy in certifyingthe controllers.

Controllers are inspected withspecial attention to their ability tosignal—that is, to notify remote per-sonnel of any abnormal conditionsin the controller system—as well asto be able to accept remote instruc-tions for starting. Diesel engine con-trollers, besides providing a startingsignal to the engine, must also pro-vide charging current for the batter-

ies, perform a weekly starting andrun test of the engine and drivenequipment train, as well as providevisual and audible indication of var-ious engine failures. These includefailure of the engine to start, shut-down from overspeed, battery fail-ure, battery charge failure and otherabnormal conditions. Diesel enginecontrollers also include pressurerecorders to sense pressure in thefire protection system and confirmunit performance on demand orduring weekly tests.

Electric motor control require-ments also focus on the ability ofthe unit to start the motor drive.This overriding concern is demon-strated in unit design, for fire sys-tem controllers are different fromstandard motor controllers. This isperhaps best demonstrated by theiruse of either a listed fire pump cir-cuit breaker or a non-thermalinstantaneous trip circuit breakerwith a separate motor overcurrentprotective device for protectionagainst overcurrents and short cir-cuits. Another difference betweenthese and standard controllers isthat once a fire pump controller isunder emergency conditions, it isprevented from shutdown exceptwhen it is under a condition morethreatening than the fire or untilconditions return to normal standby.

Surprisingly, UL lists no fire pumpcontrollers for installation in haz-ardous (classified) locations. To meetDivision I or II requirements, con-trollers must be custom designedwith a suitable hazardous locationprotection method. They must eitherhave a purge system or be explosionproof. Custom designs, of course,will incorporate either NEMA 7 com-ponents or enclose elements in explo-sion-proof boxes. The X and Z-purgesystems are the most common.

For international installations, theproblems are compounded becauseno harmonized IEC (InternationalElectrotechnical Commission) stan-dards exist that address fire pumpcontrollers This situation, however,is being addressed by a joint workingcommittee of the NFPA and UL thatis developing an IEC standard that

will incorporate NFPA 20 and UL218requirements. At present, the com-mittee has developed a draft specifi-cation and submitted it to the IECfor inclusion as a new work itemproposal (NWIP). Following accep-tance as a NWIP, the IEC will con-vene a group to review and commenton the standard for development asan IEC publication. Given the impor-tance of international markets, thismove will be an appreciated step forU.S. based manufacturers.

ConclusionFire pump systems are designed

to protect life and property. In thisarea, specifiers, owners and opera-tors need to change their perspectiveon equipment, controls and opera-tion. Performance and reliabilitymust be the priority in design, pur-chase and maintenance. While sitespecifications and purchaser require-ments certainly must be considered,simpler is usually better. The abilityof the system to perform reliablyunder the most severe conditions, forthe protection of life and property,must be the driving consideration. n

Special thanks to Kerry Bell andDave Styrcula at UnderwritersLaboratories for their assistance.

George Lingenfelder is a principaland founder of Precision PoweredProducts, a systems design and fabrica-tion firm serving the oil and gas indus-try the petrochemical/refining markets.He worked for major engineering andconstruction companies as a SystemsEngineer for ten years before startingPrecision Powered Products in 1984.Mr. Lingenfelder received his B.S.degree in mechanical engineering fromthe University of Houston

Paul Shank is with PrecisionPowered Products in Houston, Texas.He has more than 20 years of salesand engineering experience in the spe-cialty process equipment field. He hasauthored technical articles and co-authored several software programsfor the selection and pricing of pumps,compressors and steam turbines. Mr.Shank received a B.S. degree from theUniversity of Houston.

The Pump Handbook Series 255

Page 255: Centrifugal Pumps Handbook

Once the economics of a min-eral deposit have beendetermined as favorable anda decision made to proceed

with mine development, significantfinancial commitment is placed atrisk in expectation of a calculatedreturn on investment. Many aspectsof mining carry relatively high lev-els of uncertainty that contribute tothe overall degree of risk. Oneimportant consideration is ground-water control, should it be a factor,during mining operations. Wheremining must take place below thewater table in highly permeablegeologic material, operations wouldnot be possible without effectivecontrol of groundwater. Severalmines in the western United Stateshave developed large well fields, upto 70,000 gpm capacity, to interceptinflow and divert groundwater fromthe workings.

Mine Dewatering ObjectivesThere are two general objectives

to most mine dewatering programs:Keeping the working conditions rel-atively dry and maintaining the sta-bility of the excavation or opening.

Operating costs and productionrates are directly influenced byworking conditions. Wet floors andworking faces create poor groundconditions for heavy loading andhauling equipment. They alsoincrease tire wear, reduce cycletimes and impact tonnage factors.Safety becomes a direct issue if elec-

tric powered machinery is used. Inthe worst case, a submerged work-ing level becomes inaccessible andproduction is halted.

The stability of open pit wallsand underground openings is ofvital concern for safety and eco-nomic reasons. Adequate drainagemust occur in order to keep porepressure at acceptable levels basedon geotechnical stability analysis.

General Types of Mines andGroundwater Control Methods

Mines are either open pit excava-tions or underground excavations,or both. In some cases large scaleopen pit mines have succeededprior episodes of underground min-ing in the same area. In other situa-tions, underground mines are devel-oped adjacent to or from withinexisting open pit mines. Driven bymetal prices and advancing technol-ogy, companies have continued toexplore deeper and/or more chal-lenging geologic territory for mine-able orebodies. Mine dewateringmethods have evolved out of neces-sity in response to the increasinggroundwater control requirementsof contemporary mining.

Where conditions permit, openpumping from collection sumps is astandard practice. This method iscommonly used in open pit minesto control surface water drainageand in both open pit and under-ground mines where groundwaterinflow rates are small enough to be

managed in this manner. Boosterstations might be required employ-ing positive displacement pumps orhorizontal centrifugal pumpsdesigned for high head dirty water,abrasive solids or slurry service.Depending on the hydrogeologic set-ting, however, some mines cannotbe effectively or safely dewateredusing this method alone.

Well Field Systems for DewateringSeveral mines in northern

256 The Pump Handbook Series

Well Pump Applicationsfor Mine DewateringChoosing the right pumps means knowing drainage requirements, dewatering

schedules and well construction as well as system and fluid conditions.

by Mark List, Miller Sales and Engineering

Photo 1. In-pit well in service withloading operation on left, blastedmaterial on right, and high wall inbackground

CENTRIFUGAL PUMPS

HANDBOOK

Page 256: Centrifugal Pumps Handbook

Nevada require the use of wells tointercept and lower groundwaterlevels around open pit and under-ground excavations. The orebodiesassociated with these mines arehosted in fractured bedrock forma-tions along mountain ranges thatcontact alluvial basins high ingroundwater storage. Mining com-panies drill large diameter deepwells in bedrock fracture zones test-ed for favorable production yield.Other wells are completed so as tointercept shallow recharge or pro-mote drainage of less permeablezones.

Wells are typically located out-side the open pit, but drilling sitesin the pit are often unavoidable dueto local hydraulic compartmentaliza-tion. A well field can be comprisedof 30 or more individual wells withcompletion depths up to 2000 feetor greater and production casingdiameters up to 24 inches. Well-spe-cific capacities can exceed 60 gpmper foot of drawdown. Vertical tur-bine line shaft pumps (400 hp) arein service at setting depths of 1020feet, and 1500 hp vertical turbineline shaft pumps are in service atsetting depths of 800 feet. Single2200 hp submersible units work atsetting depths of more than 1800feet, and at more than 2000 feetdeep single 1500 hp submersible

units are applied. Submersibles arealso installed in series using tandem1000 hp (2000 hp) units, tandem1170 hp (2340 hp) units and combi-nation 1500 hp/800 hp (2300 hp)units with the lower setting atdepths exceeding 2000 feet.

Pit perimeter wells typically dis-charge to gathering mains at rela-tively low well head pressures. In-pit well pumps can be staged for theadditional head required to dis-charge to a surface location outsidethe pit, or to a booster station in thepit. Vertical turbine can pumps andhorizontal centrifugal pumps are inservice for this application.

Understanding the ApplicationThe uncertainties involved in

developing an efficient mine dewa-tering program become much betterunderstood as operations progress.Groundwater flow informationavailable at the onset of large scaledewatering can be very completeand supported by sophisticatedmodel simulations, but such infor-mation is usually based on field testdata that cannot be conducted at ascale proportional to what will actu-ally be undertaken. Granted, pumpapplications engineers are mostcomfortable when customersassume all risk by specifying thenecessary conditions for pumpequipment selection. The outcomeis likely to be better for all partiesinvolved if knowledge is sharedprior to establishing the conditionsfor equipment selection.

Getting the Right ConceptWell field dewatering involves

lowering the water pressure level inthe mine area to permit safe, effi-cient excavation. Over the life of themine, the change in pumping liftfrom the initial static water level tosome future pumping water levelcan be large—on the order of 1000feet or more. Individual well pro-duction capacities can decrease dra-matically depending upon aquifersystem characteristics. From apump applications perspective, thismeans selecting equipment with ini-tial operating points that best matchstarting conditions and which canbe made to fit, if possible, condi-tions that are expected to occur as

formation dewatering progresses.

Dewatering Schedule andPumping Rates

The rate at which a mine isexpected to be deepened below thestatic water level is an importantplanning factor. It is used to estab-lish a schedule for lowering ground-water heads before excavationbegins, and it is a major considera-tion in predicting the required over-all pumping rate. The change inpumping lift over time, indicated bythe dewatering schedule, is the vari-able component required to evalu-ate intermediate and final TDH con-ditions for pump selection. Pumpcapacity range can be estimatedassuming that sufficient test or oper-ating data are available to be confi-dent in doing so.

The most reliable values for indi-vidual well production capacity andefficiency (well drawdown) are notavailable until the well is construct-ed and test pumping has been com-pleted. However, the overall minedevelopment schedule might notallow for the long delivery timesthat may be needed for specialpump engineering or construction.If this potential problem is notaddressed during project planning,pump equipment orders can beplaced with results that are not cost-effective over the long term.

Well ConstructionWell dimensions limit the size

and type of pump equipment thatcan be installed. Although very cost-ly to construct, large diameter deepwells can accommodate the installa-tion of the large diameter four-pole(1800 rpm) 1500 hp and 2200 hpsubmersible electric motors that arerequired for high volume deep setapplications beyond the practicalsetting depth limitations of lineshaft pumps. Similar deep set appli-cations with smaller casing diame-ters can require the use of two-pole(3600 rpm) submersible motors withan overall length of 100 feet orlonger. In both situations a wellmust be drilled deep enough toachieve the design pump intake ele-vation and to accommodate motorequipment length and standardclearances.

The Pump Handbook Series 257

Photo 2. Twin 800 hp, 5000 gpm verticalturbine can boosters with 42” dischargemain in background

Page 257: Centrifugal Pumps Handbook

In addition to well diameter, wellalignment is of critical importancefor deep set line shaft pumps. Eventhe closest attention to constructionalignment standards, however, can-not prevent ground movement fromadversely affecting well alignmentas dewatering progresses.Depending on the severity ofground movement and resultingdeflection, shaft vibration can resultin shaft and motor bearing failure.

System ConsiderationsSystem conditions vary in

response to production well fieldchanges and discharge method mod-ifications. A well head pressure con-dition can usually be determined foruse in staging the well pump, but aconservative approach is oftentaken to ensure that the desiredpumping capacity can be main-tained. If required, throttling is usedto impose pressure temporarilyuntil system conditions are withinpump operating conditions.

Fluid Conditions Water temperature and corrosivi-

ty are major factors influencing theselection of dewatering well pumpequipment. Water temperature caninfluence the type of constructionand materials used in a line shaftpump, but elevated water tempera-ture adds significantly to the cost ofsubmersible electrical equipmentand thus can be a limiting factor inselection. At one particular dewater-ing operation, line shaft pumps arenot an option, and submersiblemotors rated at up to 2200 hp areoperating in water temperatures of140°F. These are oil filled motors ofspecialized construction sometimesfitted with heat exchangers. Becausethe motors are located in the lowerreaches of the wells, below the pumpintake, shrouds designed for adequatewater flow past the motors are usual-ly required for cooling purposes.

Unforeseen corrosion damage topump cases, impellers and columnpipe joints can ruin the best effortsin hydraulic applications engineer-ing. Corrosion potential can some-times be estimated up front bywater quality analysis, and shouldbe taken into account, if possible, inthe specification of pump equip-

ment materials and coatings.Unfortunately, geologic formationwater conditions can and do changeduring dewatering. Partial aerationcan occur with the rapid displace-ment of groundwater, and this canlead to unanticipated corrosive dam-age. Bronze and bronze alloysshould be considered if conser-vatism is justifiable. If standardmaterials are selected, the firstpump tear-down will reveal whatdoesn’t work.

Equipment SelectionMaking a reasonable attempt to

understand the factors that dictateinitial conditions and influencefuture conditions is key to selectingdewatering well pumping equip-ment that will remain effectiveunder actual operating conditions.Nevertheless, there are limitations,and well yields can eventuallydecline to the point that pumpsmust either be operated intermit-tently or replaced with lower capac-ity units.

Electric Submersible vs. Line ShaftPumps

Vertical turbine oil lubricated lineshaft pumps are operating success-fully at setting depths of more than1000 feet. The slower pump and dri-ver speeds (1800 rpm or less) ofthese units are favored by manyoperators. Pump damage from abra-sive particles or partially aerated for-mation water is significantly lessintense at slower speeds. Electricalproblems are much simpler to trou-bleshoot and correct because motorsare located at the surface above thedischarge head. On the downside,slower speeds require larger bowlsand well casing diameters. Line

shaft equipment is mechanicallycomplex and requires special engi-neering and manufacturing for bowltolerances to accommodate theeffects of relative shaft elongationunder high thrust loading. Wellalignment problems can adverselyaffect shaft and bearing life or evenpreclude the use of line shaft equip-ment.

High capacity submersible equip-ment is available in both 1800 rpmand 3500 rpm classes. Small diame-ter high yield wells or setting depthsgreater than 1000 feet generallyrestrict pump equipment to the sub-mersible type. The limitation here ismotor power output. Four-pole 20-inch diameter motors (1800 rpm) areavailable to 2200 hp for 24-inch cas-ing applications. Slim line two-polemotors (3500 rpm) can be coupled intandem to produce more than 1000hp. These two-pole motor assem-blies are more than 100 feet long,requiring additional well depth.Because the motors are installedbelow the pumps, electrical faultsthat occur in the down hole powercable or motor system requireretrieval of the equipment stringfrom the well for testing and repairto take place.

258 The Pump Handbook Series

Photo 4. Pump rig installing deep setline shaft pump

Photo 3. 400 hp vertical turbine lineshaft (1020' setting) with mineral pro-cessing plant in background

Page 258: Centrifugal Pumps Handbook

The Pump CurveFor a desired initial performance

and estimated final performance,there is a simple rule of thumb fordewatering pump selection: start onthe right side of the H-Q curve, runback to the left through the BestEfficiency Point, and plan to refitthe pump end with additional stagesif necessary to conform with esti-mated future conditions. Anothermethod is to throttle the pump dur-ing initial operations if the range ofexpected conditions indicates thatthis will eliminate the need to pulland refit the pump end. Throttlingis most common in deep set sub-mersible applications involving rela-tive certainty in the drawdown rateand final conditions.

Pump Mechanical ConsiderationsLine shaft applications involving

deep settings and high thrustrequire special consideration for rel-ative shaft stretch and bowl endplayrequirements to establish adequatelateral impeller clearance underrunning conditions. Enclosing tubetension design as well as manufac-turing tolerances for tube, shaft andcolumn pipe lengths also need to beconsidered.

Surface EquipmentPower transformers and switching

equipment are available in modularform on skid mounted platforms.Also, individual components can becustom assembled on a common skid

or placed on individual pads if pre-ferred. Mine power distribution sys-tems often are plagued with swingloads and transients depending onthe variety of electric machinery inservice. Power factor correction, sys-tem protection and motor controlrequirements vary with the applica-tion. Vertical hollow shaft typemotors used with line shaft pumpsare usually 460 volt or 4160 volt.Submersible motors are typically 460volt, 2400 volt or 4160 volt.

Installation ConsiderationsLine shaft pump installation can

be more mechanically involved thansubmersible pump installation. Theoil tube and shaft are usuallyshipped assembled in lengths of 20feet and must be individuallyplaced in each piece of column pipebefore installation in the well.Because the column, tube and shaftassembly are run in the well casing,three threaded connections must beproperly made at each 20-foot inter-val. The projection dimensions ofthe tube and shaft, which start atthe pump discharge case, must bemaintained over the length of thecolumn assembly for proper fit atthe discharge head and motor cou-pling.

Depending on the manufacturerof the submersible pump equip-ment, motor system and pumpassembly during installation can bemore or less complicated, generallyrequiring manufacturer’s field ser-

vice in addition to the installationcrew and equipment. However,once the pump and motor equip-ment are assembled in the well andtested for continuity, column instal-lation is a straightforward process ofmaking one threaded joint per pipelength and securing the power cableto the column. This goes relativelyquickly, especially if the pump rigcan handle pipe lengths of 40 feet.Operating Considerations

It is important to follow up onthe performance of a pump after ithas been placed in service. Theoperator will no doubt informsomeone associated with the saleof the equipment if a failure hasoccurred or performance is not asrepresented; conversely, the opera-tor will be concerned with othermatters if equipment performanceis acceptable. Either situationinvolves information that can assistthe applications engineer in select-ing proper equipment and recom-mending the most effective modifi-cations. n

Mark List has more than 25 yearsof experience in the construction andmining industries, with 10 years ofpractice in groundwater investigationand mine dewatering equipmentdesign, construction and operation asChief Dewatering Engineer for a west-ern gold mine. He is a registered engi-neer and state water rights surveyor inNevada.

The Pump Handbook Series 259

Page 259: Centrifugal Pumps Handbook