The Relatioship of Vibration to Problems in Centrifugal Pumps - Che

Pump reliability problems can beresponsible for a large amountof the maintenance budget andlost-opportunity cost at chemi-

cal process plants. Typical reasonsfor pump failures can be diagnosedearly by applying the right kinds ofvibration testing, analysis, and eval-uation criteria during pump monitor-ing or troubleshooting.

The most important mechanical andhydraulic issues relevant to reliablepump operation include level of imbal-ance, level of misalignment, where thepump is operating on its curve [prefer-ably close to best-efficiency point(BEP)], how its net positive-suction-head required (NPSHR) compares tothe worst-case suction head available,the configuration of the piping hy-draulics in the suction piping close tothe pump, and the manner in whichpiping suction and discharge nozzleloads are accommodated. Problemswith any of these issues show up assymptoms, which include higher thannormal vibration at certain key fre-quencies. Often, a simple plot of vibra-tion-level versus frequency “spectrum”is sufficient to allow a diagnosis. How-ever, when actions based on such in-formation do not resolve the situation,more detailed vibration testing shouldbe considered, such as operating de-flection shape plotting and modal-analysis “bump” testing.

After discussing the reasons forthe importance of the listed me-chanical and hydraulic issues, wewill use several plant examples todescribe how to apply various levelsof vibration testing in pump evalua-

tion, and in the determination of ef-fective fix options.

Design & selection influencesA pump should be selected that willtypically operate close to its BEP. Con-trary to intuition, centrifugal pumps donot undergo less vibration as they arethrottled back, unless the throttling isaccomplished by variable-speed opera-tion. Operation well below the BEP atany given speed, just like operationwell above that point, causes a mis-match in flow-incidence angles in theimpeller vanes and the diffuser vanesor volute tongues of the various stages.This loads up the vanes, and may evenlead to airfoil stalling, with associatedformation of strong vortices (miniaturetornadoes) that can severely shake theentire rotor system, and can even leadto fatigue of impeller shrouds or dif-fuser plates or “strong-backs.” Rotor-impeller, steady-side-loads and shak-ing that occurs at flows below the onsetof suction or discharge recirculationleads to potential rubs, and excessiverotor loads that can damage bearings.

Many plants buy equipment that hasmore capacity than is needed, to allowfor future production expansion. But indoing so, they ensure years of unreliableperformance of potentially reliable ma-chinery. The typical effect on vibration ofthis practice is shown in Figure 1. If youmust run a pump away from its BEP be-cause of an emergency situation, planteconomics, or other operational con-straints, at least never run a pump forextended periods at flows below the“minimum-continuous flow” provided bythe manufacturer. Also, if this flow was

specified prior to about 1985, it may bebased only on avoidance of flashing andnot on recirculation onset, and should berechecked with the manufacturer.

System design & installationSuction design. The design of thepump suction has significant mechan-ical ramifications. Both the mechani-cal connection of the suction flangeand the hydraulic design upstream ofthe pump impeller are of key impor-tance in this regard.

Relative to the mechanical connec-tion, avoid unrestrained expansionjoints (piping “flexible joints”) at largenozzles. The pressure across the cross-section of such nozzles, multiplied bythat area, becomes a large thrust, sim-ilar to the thrust at the exit of a rocketnozzle. Just because the casing is notfree to move does not mean that it isnot distorted by the thrust at the unre-strained nozzle; in fact, it makes thecasing more likely to distort. Thisthrust typically does not overstress thenozzle, but the excessive distortionthat it causes in the casing or base-plate leads to severe operating align-ment problems and possible rubs.

The main issue of hydraulic concernis that sufficient suction-static pres-sure be present to avoid cavitation.Today, it is understood that thismeans more than merely having suffi-cient net positive-suction-head avail-able (NPSHA) to satisfy the 3% headdrop NPSHR published by the manu-facturer.

At NPSHA, as much as three timesthe NPSHR, high-frequency cavita-tion (sometimes inaudible) can causeserious erosion of the suction side ofimpeller vanes or wear-ring exits.Even if NPSHA is high enough toavoid cavitation under normal circum-stances, this cavitation can still be

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38 CHEMICAL ENGINEERING WWW.CHE.COM MAY 2004

THE RELATIONSHIP OF VIBRATION TO PROBLEMS IN

Centrifugal Pumps William D. Marscher, P.E.

Mechanical Solutions, Inc.

Vibration analysis, when properly carried out, can help keep your pumps operating troublefree

FIGURE 1. As flow is reduced from a pump’sBEP value, vibrations increase significantly,

more or less linearly, until the onset of‘suction-recirculation flow.’ At lower flows

than this, the vibration gets much higher, andbecomes erratic and uncertain, as shown by

the cross-hatched ared

caused in local sectors by skewed orswirling flow in the inlet pipe, as canbe the result of an elbow too close tothe pump-suction flange, or by too se-vere a reducer near the suction flange,or by vortices in the inlet sump.

If the pump is operated too far awayfrom its BEP, the angle of attack of in-coming flow on the rotating impellervane can be different from that antici-pated by the pump designer at thatpump speed, and vane stalling canoccur at either the suction or discharge,leading to suction or discharge recircu-lation, respectively. Such internal flowrecirculation can cause cavitation onthe pressure-side of vanes, and cancause tornado-like eddies that rotatewith the impeller, but at a somewhatslower speed, exciting rotor-criticalspeeds at unexpected frequencies.Beware of VFD resonance. Vari-able-frequency drives (VFDs) save en-ergy by allowing a pump’s flow to bereduced without throttling the dis-charge. This comes with an unex-pected side-effect, however, of increas-ing the chance that excitationstypically strong at 1x the pump’s run-ning speed, such as residual imbal-ance, will excite a natural frequencyinto resonance, causing unacceptablyhigh vibration levels.

The natural frequency is the num-ber of cycles per minute at which therotor or structure will vibrate if it is“rapped,” like a tuning fork. Pump ro-tors and casings have many naturalfrequencies, some of which are gener-ally in, or close to, the operatingspeed range. The vibrating patternswhich result when a natural fre-quency is close to the running speedor some other strong force’s frequencyare known as mode shapes. Each nat-ural frequency has a different modeshape associated with it. Knowing

this shape, for example through de-tailed vibration testing as discussedbelow, provides insight as to wheremaximum leverage can be placed tostiffen the natural frequency out ofthe range of the resonant excitation.

In resonance, the oscillating energyfrom the last “hit” of the force comesfull cycle, and is then reinforced whenthe next hit takes place. The vibrationin the next cycle will then includemovement due to both hits, and willbe higher than it would be for one hitalone. The vibration motion keepsbeing amplified in this way until itsconsiderable motion uses up more en-ergy than the amount of energy that isbeing supplied by each hit. Unfortu-nately, the motion at this point is gen-erally large and damaging.

Resonance is illustrated in Figure 2.You want the natural frequencies ofyour rotor and bearing housings to bewell separated from the frequencies thatresonant “dribbling” type forces willoccur at, which tend to be 1x runningspeed (typical of imbalance), 2x runningspeed (typical of misalignment), or at thenumber of impeller vanes time runningspeed (so-called vane-pass vibrationsfrom discharge-pressure pulses as theimpeller vanes move past the volute ordiffuser vane “cut-water”).

Certain types of nonlinear-vibrationresponse have been found common inturbomachinery. They generally fallinto the category of parametric reso-nance, and are beyond the scope of thisdiscussion. They can result in large vi-bration in spite of relatively low drivingforce. Typically, such resonances arecaused by bearing support looseness ora rub at a bearing, seal or other runningclearance. The symptoms are a pulsat-ing orbit, with a large amount of vibra-tion at exact whole fractions of runningspeed, such as 1/2 or 1/4.

Besides sweeping the excita-tion frequencies through alarge range, and therefore in-creasing the chance of match-ing running speed to one ofthese natural frequencies,many (especially inexpensiveor older) VFD controllers pro-vide new excitations at various“control-pulse” multiples of themotor-running speed, com-monly at 6x and 12x, and often

at whole-fraction submultiples as well.Balance. Imbalance is the most com-mon cause of excessive vibration in ma-chinery, followed closely by misalign-ment. Balance is typically thought of asstatic (involving the center of massbeing off-center so that the principal-axis-of-mass distribution is still parallelto the rotational centerline) and dy-namic (the principal-mass axis makesan angle with the rotational axis). Foraxially short components (such as athrust washer), the difference betweenthese two can be neglected, and only sin-gle-plane static balancing is required.For components greater in length thanone fourth their diameter, dynamic im-balance should be assumed, and at leasttwo-plane balancing is required.

When imbalance occurs, including im-balance caused by shaft bow, its showsup with a frequency of exactly 1x runningspeed. This is because the heavy side ofthe rotor is rotating at exactly rotatingspeed, and so forces vibration movementat exactly this frequency. Typically, thisalso leads to a circular shaft orbit, al-though the orbit may be oval if the rotoris highly loaded within a journal bearing.This is shown in Figure 3.Pump and driver alignment. Nextto imbalance, misalignment is themost common cause of vibration prob-lems in rotating machinery. Misalign-ment is usually distinguished by twoforms: offset, and angular. Offset isthe amount that the two centerlinesare offset from each other (that is, thedistance between the centerlineswhen extended sufficiently to be nextto each other). Angular misalignmentis the differential crossing angle thatthe two shaft centerlines make whenprojected into each other, whenviewed first from the top, and then ina separate evaluation from the side.In general, misalignment consists of a

CHEMICAL ENGINEERING WWW.CHE.COM MAY 2004 39

FIGURE 2. Illustration of natural-frequencyresonance, and the effects of damping

FIGURE 3. Typical vibrationspectrum and orbit data due toimbalance

combination of both offset and angu-lar misalignment.

Offset misalignment requires eithera uniform horizontal shift or a consis-tent vertical shimming of all feet of ei-ther the pump or its driver. Angularmisalignment requires a horizontalshift of only one end of one of the ma-chines, or a vertical shimming of justthe front or rear set of feet. Combinedoffset and angular misalignment re-quires shimming and/or horizontalmovement of four of the combinedeight feet of the pump and its driver.In principle, shimming and/or hori-zontal shifting of four feet only shouldbe sufficient to cure a misalignment.

Sometimes a rotor must be offsetwhen cold and not running in order to bealigned when running and hot. For ex-ample, thermal expansion of the pumpor its pedestal versus the driver couldchange alignment “hot” versus “cold.”This situation can be checked with vari-ous continuous monitoring tools, such asDodd Bars or Essinger Balls or Bars,which monitor the offset and angularmovement of the pump and driver bear-ing housings relative to either each other(Dodd Bars) or relative to the foundationas a reference (Essinger Balls).

When misalignment is a problem, ittypically causes primarily vibrations at2x running speed, because of the highlyelliptical orbit that it forces the shaft torun in on the misaligned side. Some-times the misalignment load can causehigher harmonics (that is, integralmultiples of the rotor speed; especially3x), and may even decrease vibration,because it loads the rotor unnaturallyhard against its bearing shell.

Alternately, misalignment may ac-tually cause increased 1x vibration, bylifting the rotor out of its gravity-loaded “bearing pocket,” to result inthe bearing running relatively un-loaded. This can also cause shaft in-stability, as discussed later.

However, Figure 4 shows a typicalorbit and fast-Fourier transform (FFT)spectrum for misalignment, in which 2xrunning speed is the dominant effect.This is often accompanied by relativelylarge axial motion, also at 2x, becausethe coupling experiences a non-linear“crimp” twice per revolution.

Misalignment is best checked bysome form of the reverse-dial-indica-

tor method (or its laser-alignmenttool equivalent).Rotor dynamics. Dynamics of sta-tionary components, such as casing-pedestal assemblies, is a straightfor-ward extension of linear-elastic-staticanalysis, and is usually performed byusing simplified formulas such asthose of Blevins [4], or by using a gen-eral purpose finite element program.In contrast, rotor dynamics requires amore specialized computer program,which deals with effects such as:• Three-dimensional stiffness and

damping at bearings, impellers andseals as a function of speed and load

• Impeller and thrust-balance-devicefluid excitation forces (includingcross-coupling as described below)

• Mass added by the pump fluid• Gyroscopic effects• Lomakin effect (explained below)

The Lomakin effectPump wear rings and (in multistagepumps) interstage bushings and bal-ancing devices can develop large forcesin their annular-clearance space dueto the circumferentially varyingBernoulli pressure drop induced asrotor eccentricity develops. This isknown as the Lomakin effect, and candramatically affect pump rotor dynam-ics, especially in two ways:• By changing the rotor-support stiff-

ness and therefore the rotor naturalfrequencies, thereby either avoiding orinducing possible resonance betweenstrong forcing frequencies at one andtwo times the running speed and oneof the lower natural frequencies

• By increasing the rotor-supportcross-coupled stiffness or, as theseals wear, decreasing the rotordamping, allowing greater magnifi-cation factors during resonance,and, in the extreme, allowing desta-bilizing forces on the rotor to domi-nate those associated with stabiliz-ing energy absorption

These effects are particularly strongin multistage pumps, because multi-stage rotors are relatively long andflexible. Since annular seals (such aswear rings) are primarily in the cen-tral portion of the rotor, where theyexercise considerable leverage on thefirst bending mode of the rotor, thecontribution of seal stiffness to the

rotor support can be comparable tothat of the rotor stiffness itself.

The Lomakin effect depends di-rectly on leakage and therefore thepressure drop across the seal, which,for constant-system flow, results in avariation of the Lomakin support stiff-ness with roughly the square of therunning speed. However, the moreusual case of roughly constant systemhead results in only a small variationof Lomakin effect with pump speed.The other important parameters areannular sealing cavity length, diame-ter, and clearance. Fluid propertiesare of secondary importance.

The radial-clearance effect is astrong influence, with Lomakin effectroughly proportional to its reciprocal.The physical reason for the strong in-fluence of clearance is that it gives theopportunity for the circumferentialpressure distribution, which is behindthe Lomakin effect, to diminishthrough circumferential flow. Any sealsurface which includes grooving has asimilar effect as increased clearance.

The primary action of the Lomakineffect is beneficial, through increasedsystem direct stiffness and dampingtending to increase the rotor naturalfrequency and decrease the rotor vi-bration response at that natural fre-quency. However, over-reliance on Lo-makin effect can put the rotor designin the position of being too sensitive towear of operating clearances, resultingin unexpected rotor failures due to res-onance. It is important that modernrotors be designed with stiff enoughshafts, so that any natural frequencythat starts above running speed withnew clearances remains above the run-ning speed with clearances worn to thepoint that they must be replaced froma performance standpoint.

Testing for difficult diagnosesExperimental modal analysis is amethod of vibration testing in which aknown force (constant at all frequen-cies within the test range) is put into apump, and the pump’s vibration re-sponse exclusively due to this force isobserved and analyzed. The actualnatural frequencies of combined cas-ing, piping, and supporting structurecan be accurately determined.

Figure 5 shows how the broadband

Cover Story


“terrain” or peaks and valleys from amodal-impact test can be used to betterunderstand whether the high vibrationsare caused by the amplifying effect of anatural frequency, or are just due to anexcessive-oscillating force. Knowing thisallows you to focus on the root cause,and provide a fundamental fix.

The main tools required to do modaltesting are a multichannel FFT fre-quency analyzer, a PC with commer-cially available software, a set of vibra-tion-response probes (such asaccelerometers), and an impact ham-mer designed to spread its force over afrequency range that covers the testrange. The impact hammer has an ac-celerometer in its head, which is cali-brated to indicate the force being ap-plied. During a modal test, the signalfrom the hammer input force ac-celerometer is sent to one channel ofthe spectrum analyzer, and the signalfrom the vibration response probe issent to the second channel. Dividing, ateach frequency, the second channel bythe first channel gives the frequency-response function (FRF) of the pumpand its attached system. The peaks ofthe FRF are the non-critically dampednatural frequencies, and the width andheight of the peaks indicate the damp-ing of each natural frequency, and howsensitive vibration at the test locationis to forces which occur in the vicinityof the hammer impact at frequenciesnear a given natural frequency.

Modal tests determine the funda-mental source and general nature of apumping system’s vibrations, and pre-dict the effect of system modifications.A useful variation of modal testing hasbeen developed by the author toachieve this within the time and opera-tional constraints of actual field tests.

This technique is known as time-aver-aged pulse (TAP), as described in thereferences, and provides the benefitthat the pump does not need to be shutdown in order to perform the “bump”test. The TAP method uses statistics onseveral hundred impacts to filter thedata obtained from modal analysis inorder to reliably determine structuralnatural frequencies and mode shapes,resonating force locations and frequen-cies, and rotor critical speeds while thepump is operating at the problem con-dition. TAP then uses classical modal-analysis-processing techniques to pro-duce an animated model of eachnatural frequency’s vibration pattern,and to make predictions about the ef-fectiveness of proposed design changes,such as stiffer bearings, new pipingsupports, or a thicker baseplate. TheTAP method may be applied to ma-chines at any operating speed and load.

TAP testing can result in a compli-cated modal-test data base consistingof frequency plots at many locations ofvibration response to the calibratedhammer impact. Computer programsare available from various testing soft-ware companies that will plot a scale-up of that data to provide animations ofthe vibration mode shape. In some ofthese computer programs, this infor-mation can be used to automaticallypredict the best locations for addedmasses, dampers or stiffeners to solvethe vibration problem associated witha given mode. Similar cartoons can alsobe made of the operational vibrationtaken at many locations and in manydirections, and, the result is known asa operating deflection shape (ODS).

The TAP procedure can be used todetermine the critical speeds of apump rotor while the pump is operat-ing at the condition of interest. Thisapproach provides an alternative to

the current practice of basing criticalspeeds on the peaks of “waterfallplots” of spectra obtained duringpump startups or shutdowns. Water-fall plots may represent pump-rotor-dynamic performance poorly duringsteady-state operation due to thestrong sensitivity of annular-seal stiff-ness and damping to operating condi-tions. The primary reason that pumpnatural frequencies are sensitive tooperating condition is the Lomakin ef-fect. As discussed earlier, the Lo-makin effect can significantly changethe rotor-support stiffness and damp-ing, depending upon the pump’s speedand load, making “run-down” water-fall plot testing of uncertain value.

Recognize and solve problemsProbably 90% of all pump vibrationproblems can be solved by careful bal-ancing of the rotor assembly, alignmentof the coupling when the system is at itsrated conditions (especially if it is hot),for example per Dodd, and running thepump within the bounds of its specifiedhead versus capacity curve. Remainingpump-vibration problems are generallydue to a resonance of a system naturalfrequency with one of the excitationforces common to all pumps, such asresidual unbalance. During resonance,the rotor vibrations can exceed clear-ances, or excessive bearing loads canoccur, even if unbalance, misalignment,and hydraulic loads are within nor-mally acceptable limits.

The following measurements aresuggested as a minimum for predic-tive maintenance or trouble-shootingof any style pump:• What the vibration level is on both

bearing housings on pump, and onpump-side (that is, “inboard”) bear-ing housing on driver in the vertical,horizontal and axial directions

• How hydraulic performance comparesto design. In other words, for a givenspeed and capacity (that is, flowrate),how close is the temperature-compen-sated head of the pump to the curvesupplied by the manufacturer, espe-cially near the design or BEP? Are thehead and capacity steady when theoperator tries to hold the pump at aconstant speed? Is the motor or steamturbine driver required to providemore power than expected?


FIGURE 4. Misalignment example of shaft orbit and FFT spectrum

FIGURE 5. Vibration frequencyspectrum dependence of naturalfrequencies and forces (F)


TABLE 1. VIBRATION FREQUENCIES VERSUS TYPICAL CAUSESMultiples of Running Speed: Other Symptoms: Probable Causes:0.1 to 0.3 X Vibration response is broad in frequency Diffuser or return channel stallExactly 1/2 X, Vibration may increase dramatically shortly after this appears, except in Light rub, combined with low 1/3 X, 1/4 X etc. vertical pumps with four or more lineshaft bearings, where it is common shaft or bearing support

and not harmful natural frequency0.40 X (Approx.) Shaft has rolling element bearings Bearing cage defect 0.42 to 0.48 X Near shaft natural frequency and orbit "pulses," forming an inside loop. Rotor dynamic instability due to

Vibration onset is sudden at a speed roughly 2x the excited natural fluid whirl in close clearances, frequency, and "locks" onto the natural frequency in spite of speed increase e.g. bearing "oil whip"

f= 0.6 to 0.93 X Smaller peaks at (1-f) X, and at +(1+f) X "sidebands" of the first several Internal flow recirculation multiples of running speed. Often accompanied by rumbling noise and probably at suctionbeating. Occurs at part-load capacities, but disappears at very low capacity. Independently depends on speed and flow

Less than 1 X Increased broadband vibration and noise level below running speed as Cavitation without NPSH decreases, especially at high flows. Often accompanied by recirculationdecreased vibration and noise above 1 X and increase at f >10KHz

1 X 1. Stronger on shaft than on housing. Hydraulic performance and/or suction Imbalance in rotating pressure normal. Axial vibrations within normal limits and vibrations increase assemblywith roughly speed squareda. Vibrations highest on drive IB Pump coupling imbalanceb. Vibrations highest on drive IB housing Driver rotor imbalancec. Vibration high on pump IB or OB housing, low on driver Pump rotor imbalanced. Natural frequency near 1 X Resonance

2. Axial vibrations are over 1/2 of H or V vibrations, or vibrations Pump/driver misalignment increase much slower than the square of the speed. at the couplingAlso, bearing oil temperature is high

3. Discharge pressure pulsations are strong at 1 X but not at Clogged or damaged impeller vane pass in a single volute pump impeller passage

4. Same as #3, but vane pass also strong, especially flows far Volute tongue designed too above or below the design point close to impeller OD, or exces-

sive impeller/volute eccentricity2 X 1. Axial vibrations are low

a. Both shaft and housing vibrations are strong, and discharge pressure Clogged or damaged pulsations are strong but impeller vane pass vibrations impeller passageare low in a twin volute pump

b. Same as (a) but combined with unusually high vane Volute vanes designed too close to pass vibrations and discharge pressure pulsations impeller OD, or clogged or dam-

aged volute, or excessive impeller/volute eccentricity

c. Shaft vibrations much stronger than housing vibrations, and approach Looseness in bearing or exceed bearing clearance. Decrease in shaft first bending natural support or cracked shaftfrequency. Other multiples of running speed may be stronger than usual.

d. Shaft vibrations stronger than housing and driver torque pulses Torsional excitatione. In motor-driven pumps, with speed equal to electrical Electrical problem

line frequency, and highest vibration at motor IB housing with motor2. Axial vibrations are over 1/2 of horizontal vibrations and 1 X Pump/driver misalignment

vibrations are also high and bearing oil temperature is high at the couplingNumber of impeller 1. Discharge pressure pulsations reasonably low and both Volute too close to impeller vanes times shaft and housing vibrations high OD due to design or excessive running speed rotor eccentricity

2. Same as (1) but housing vibrations much higher Piping mechanical than shaft vibrations resonance at vanepass

3. Discharge pressure pulsations high at vane pass frequency Acoustic resonance high but suction pressure pulsations reasonably low in discharge pipe

4. Suction pressure pulsations high Acoustic resonance in suction pipe5. Pump rotor or casing natural frequency close to vane pass Resonance

Several multiples of 1. Orbit shows sharp angles or shows evidence of "ringing", Internal rub or poorly lubricated running speed, in- and/or spectrum shows evidence of exactly 1/2 or 1/3 X gear couplingcluding 1X, 2X, 3X, response. Grinding noises and speed changes may be evident4X and possibly 2. Orbit is "fuzzy" but does not pulse or "ring". Seal coolant flow Jammed, clogged or higher is unexpectedly high or low, and may exhibit high temperature. damaged seal

Spectrum may also exhibit 1/3 or 1/2 running speed3. Orbit pulses, usually in one direction much more than Shaft support looseness, espe-

the other and shaft vibrates more than housing cially in bearing insert or cap reten-tion.

4. Shaft and orbit fairly steady, and housing vibrates more than shaft. Often Looseness in pump combined with vibration response over a broad range of frequencies casing, pedestal or foundation.below running speed.

Number of bearing Freq f =0.4 Nb*N/2 approx Defect bearing outer race rolling elements Freq f= 0.5 Nb*N/2 approx Defect in ball or roller(Nb) times about 1/2 Freq f= 0.6 Nb*N/2 approx Defect in inner racethe rotor speed (N)

• Whether or not the suction pressureis well above NPSH requirements

• What the bearing-shell or lubricant-exit or sump temperatures are, atleast approximately

In deciding what action to take on thebasis of the above measurements, a trou-bleshooting table can serve as a guide onhow to interpret symptoms. In Table 1,such a list is provided. It is not meant tobe all-inclusive, and is in the order of thefrequency value observed, not in order oflikelihood or importance to reliability. Inaddition, persistent pump-vibrationproblems are usually due to an unex-pected combination of factors, like me-chanical or acoustical-piping reso-nances, or hot running misalignment ofthe pump-driver due to thermal distor-tions of the piping or baseplate. Defini-tions useful in reading this list are:Runout: False vibrations picked up by aproximity probe, actually reflecting shaftscratches, gouges, undulations, and soon, or static-shaft bends or misalign-ment (mechanical runout), or eddy-cur-rent-sensitivity variations along theshaft surface or instrumentation noise(electrical runout). This can be deter-mined by observing the apparent shaftorbit when the shaft is slowly turned.Inboard: The coupling side of thepump or driver.Outboard: The end of the pump or dri-ver opposite to the coupling.Narrowband: Vibration responsewhich is at a single frequency.Broadband: Vibration responsewhich covers a band of frequency ona spectrum plot.

Information that is typically ob-tained prior to using such a list is pro-

vided in Figure 4 for the example caseof misalignment. If imbalance werethe problem, the orbit would be circu-lar instead, and the FFT spectrumwould have a high narrow peak pre-dominantly at 1x running speed.

Figure 6 shows typical failuremodes and associated frequencies in arepresentative multistage pump. Thepoint of this figure is not to be all in-clusive of pump problems, but to showmajor issues as discussed in Table 1,and how the problem relates to the re-sulting vibration.

Some examplesMultistage pump changed frombaseload to cycling service. A planthad experienced chronic boiler-feed-pump failures for eight years, sincethe unit involved had been switchedfrom base load to modulated load. Thelongest that the turbine-driven pumphad been able to last between majorrotor-element overhauls was fivemonths. The worst wear was seen tooccur on the inboard side of the pump.The turbine was not being damaged.The pump vendor had decided on thebasis of detailed vibration-signaturetesting and subsequent hydraulicanalysis that the internals of the pumpwere not well enough matched to part-load operation, and proposed replace-ment of the rotor element with a newcustom-engineered design, at a verysubstantial cost. Although the problemshowed some characteristics of a criti-cal speed, both the vendor and theplant were sure that this could not beproblem, because a standard rotor-dy-namics analysis showed that the factor

of safety between run-ning speed and thepredicted rotor criti-cal speeds was over afactor of two.

However, the finan-cial risk associatedwith having blindfaith in the hydraulicsand rotor-dynamicanalyses was consid-erable. In terms ofvendor compensation

for the design, and the plant mainte-nance and operational costs associatedwith new design installation, the com-bined financial exposure of the vendorand the plant was about $350,000.

Impact-vibration testing using theTAP cumulative time-averaging im-pact-test procedure quickly determinedthat one of the rotor-critical speeds wasfar from where it was predicted to beover the speed range of interest, asshown in Figure 7, and in fact haddropped into the running speed range.Further testing indicated that this crit-ical speed appeared to be the sole causeof the pump’s reliability problems.“What-if” iterations using a rotor dy-namic computer model showed that theparticular rotor natural frequencyvalue and rotor mode deflection shapecould best be explained by improper op-eration of the driven-end bearing. Thebearing was removed and thoroughlyinspected, and was found to have a crit-ical clearance far from the intendedvalue, because of a drafting mistake onthe bearing’s OEM drawing, which wascarried over each time the bearing wasrepaired or replaced. Installation of thecorrectly constructed bearing resultedin the problem rotor critical speedshifting to close to its expected value,well out of the operating speed range.The pump has since run for years with-out need for overhaul.

Single-stage, double-suctionpump with nozzle load-inducedmisalignment. A large double-suc-tion, single-stage pump, with an im-peller diameter of 4 ft (over 1 m) and arunning speed of 600 rpm, was de-signed with close impeller-vane/volute-


FIGURE 6. Horizontalpump symptoms oftypical hydraulic ormechanical problems,and some typicalvibrations associatedwith pump-failure modes

tongue clearance to reach an aggres-sive efficiency level in a petrochemicalfacility where energy was at a pre-mium. During installation, it wasfound that vibration amplitudes rose toas high as the operating clearances inthe wearing rings (25 mils, or 0.6 mm,diametral), with the primary compo-nent at running speed. There was nopossibility of a resonance in this pumpsince both the shaft and the bearinghousing natural frequencies wereabove the 1x and 2x excitations. Thevane pass frequency of 4,200cycles/min was far removed from theshaft first and second non-criticallydamped natural frequencies of 2,850and 19,000 cycles/min, respectively.

The reason for the high vibration wasfound to be 35 mils of misalignment atthe coupling due to the hydraulic loadson the pump discharge flange being far inexcess of API 610 recommended levels.The 48 in. discharge had a piping-expan-

sion joint at the flange, with no tie-bars in place across the flange to carrythe resulting thrust. After removal ofthe piping forces through a groundedbulkhead bolted to the dischargeflange, the pump’s large 1x and 2x vi-bration levels were reduced to accept-able values per API-610.Vertical pump with hollow driveshaft/gear-box drive. A major U.S.

petroleum refinery had a serious prob-lem of gear box failure, coupled with asevere, high-pitched noise in violation ofOSHA standards, in some service waterpumps. These pumps were driven atvariable speeds by a steam turbinethough a right-angle, 1:1 gear box andhollow-drive shafting. Many expertsfrom the pump, turbine, and gear manu-facturers, and from independent con-sulting firms, had tried unsuccessfully touse vibration signature testing, andsometimes finite element analysis(FEA), to understand and cure the prob-lem over the several years since installa-tion. Replacement of the gear boxes withones carefully built to more stringent tol-erances had no effect. It was suspectedthat the problem involved a torsionalcritical speed, excited by gear-meshingfrequency. However, torsional testingperformed by the author found that allrotor-system torsional natural frequen-cies were close to their predicted values,

and were not near the gear mesh of theunit’s single-operating speed.

Impact-modal testing was performedon all exposed stationary as well as ro-tating components, using the cumula-tive time-averaging method discussedabove. None of the results indicated thepresence of any natural frequencies closeto the excited gear meshing frequency,until the 4-ft-long, hollow drive shaftwas impact tested while operating. Thesurprising test results showed that thisshaft, when under torque, had a “bell-mode” almost exactly at the gear mesh-ing frequency. The mode shape of the ex-cited natural frequency (Figure 8) wassuch that the hollow shaft ovalized withvery little damping, causing the shaftlength to oscillate as the cross-sectioncyclically ovalized. Subsequent analysisshowed that the unexpected axial move-ment was through the Poisson effect,which states that as you strain a compo-nent in one direction, it automaticallydeflects at the same time in the perpen-dicular direction. The driving force wasshown by further testing to be the com-bined torsional and axial load from thebull/pinion gear meshing. The driveshaft was filled with grease to damp outthis unusual vibration. The gearboxnoise immediately fell a factor of ten,and all gear box problems ceased. n

Edited by Gerald Ondrey

FIGURE 8. Shapeof hollow-drive

shaft mode nearthe gear mesh

frequency

FIGURE 7.Variation of rotor-critical speedswith IB bearing stiffness

AuthorWilliam D. Marscher istechnical director of the rotat-ing-machinery consulting andtroubleshooting firm Mechan-ical Solutions, Inc. (1719Route 10 East, Suite 305,Parsippany, N.J. 07054-4507;Phone: 973-362-9920; Fax:973-362-9919; email: [email protected]). His34-year career has included de-sign, analysis, and vibration

testing for GE, Honeywell, Pratt & Whitney, andWorthington/Dresser. He was 1999 STLE president,is an STLE Fellow, is associate editor of TribologyTransactions, and chairs the STLE Seals TechnicalCommittee and ASME Predictive MaintenanceCommittee. He organized the 1993 ASME RotatingMachinery Conference and 1995 ASME Interna-tional Tribology Conference. He has written 65 pa-pers and 7 handbook chapters, is coauthor of “Cen-trifugal Pump Design & Performance,” and is on theboard of Pumps & Systems. He is a voting member oftwo ASTM committees and the ISO machinery noiseand vibration standards committee. He received aNASA fellowship, ASLE’s Hodson Award and theDresser Engineering Medal. A registered profes-sional engineer, he is on the Texas A&M Pump Sym-posium Advisory Committee, and the TAMU DubaiPump & Turbo Symposium Advisory Committee. Heholds engineering degrees from Cornell and RPI.

References1. API 610, 9th Ed., Amer. Pet. Inst., Washing-

ton D.C., 2002.2. Black, H.F., Effects of Fluid-Filled Clearance

Spaces on Centrifugal Pump Vibrations. 8thTurbomachinery Symposium, Texas A&MUniv., 1979.

3. Blevins, R.D., “Formulas for Natural Fre-quency and Mode Shape,” Robert KriegerPublishing Co., Malabar, Fla., 1984.

4. Bowman, D., others, Pump Rotor CriticalSpeeds: Diagnosis and Solutions, Proc. Inter-national Pump Symposium, Texas A&MUniversity, April 1990.

5. Childs, D., Finite Length Solutions for Rotor-dynamic Coefficients of Turbulent AnnularSeals, ASME/ASLE Lubrication Conf.,ASME Paper 82-LUB-42, Oct. 1982.

6. Dodd, V.R., “Total Alignment,” The Petro-leum Publishing Co., Tulsa, Oklahoma, 1974.

7. Ewins, D.J., “Modal Testing: Theory and Prac-tice,” Research Studies Press, Wiley NY, 1984.

8. Fraser, W.H., “Centrifugal Pump HydraulicPerformance and Diagnostics, Pump Hand-book,” McGraw-Hill, New York, 1985.

9. Jen,C.-W., and Marscher, W. Experimental ModalAnalysis of Turbomachinery Rotors Using Time-Averaging, Proc. International Modal AnalysisConference, Orlando Florida, Jan. 26, 1990.

10. Marscher, W.D., The Effect of Fluid Forcesat Various Operation Conditions on the Vi-

brations of Vertical Turbine Pumps, Proc.IMechE, Radial Loads and Axial Thrusts onPumps, Feb. 5, 1986.

11. Marscher, W.D., Determination of Pump RotorCritical Speeds During Operation through Useof Modal Analysis, Proc ASME 1986 WAMSymposium on Troubleshooting Methods andTechnology, Anaheim Calif., Dec. 1986.

12. Marscher, W., The Relationship BetweenPump Rotor System Tribology and Appropri-ate Vibration Specifications for CentrifugalPumps, Proc IMechE 3rd European Con-gress on Fluid Machinery for the Oil andPetrochemical Industries, The Hague,Netherlands, May 1987.

13. Marscher, W., Analysis and Test of Multi-stage Pump Wet Critical Speeds,ASLE/ASME Joint Tribology Conf., Ft.Lauderdale, Oct. 1989.

14. Marscher, W., The Effect of Variable Fre-quency Drives on Vibration Problems in Ver-tical Pumps, Proc. Water & WastewaterConference, Barcelona, Spain, April 1990.

15. Marscher, W., The Determination of RotorCritical Speeds While Machinery RemainsOperating through Use of Impact Testing,IMAC Conf. Orlando Fla., SEM, Feb. 1999.

16. Marscher, W., Avoiding Failures in CentrifugalPumps, Tutorial in Proc. TAMU InternationalPump Symposium, Houston Tex., Feb. 2002.

17. Rathbone, T. Vibration Tolerance. PowerPlant Engineering, Vol.43, Nov. 1939.


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