Engineering Manual.pdf

1PROPRIETARY information of ROBINSON INDUSTRIES, INC.Do NOT copy or distribute.

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Performance Specifications


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Given: Airflow = 30,000 SCFM (@70° F and density = 0.075 lb/ft^3)Air temperature from process = 200° FDischarge temperature from RTO = 400° FFD fan pressure required = 20 in-H2O

#1. Forced Draft Fan: Specify fan performance of 37,500 ACFM, 20 in-H2O (-1 in-H2O @ Inlet), 200° F, 0.060 lb/ft^3

Effect of Fan Location in System (F.D. Fans vs. I.D. Fans)

Process200°F 200°F

-1 in-H2O +19 in-H2O 0 in-H2O

400°F 400°FHeat

RTO Stack

54321

RFD Fan

Fan ∆P = 20 in-H2O RTO ∆P = 14.5 in-H2O

#2. Induced Draft Fan: Specify fan performance of 50,562 ACFM, 14.8 in-H2O(-14.8 in-H2O @ Inlet), 400° F, 0.0445 lb/ft^3

Process200°F 200°F

-1 in-H2O-1 in-H2O 0 in-H2O

400°F400°F

Heat

RTO

5321

RID Fan

Fan ∆P =14.8 in-H2O

RTO ∆P =14.0 in-H2O

Stack

400°F

4

Q P(gage) Density T(°F)1. 37500 -1 0.060 2002. 35714 +19 0.063 2003. 46875 +14.5 0.048 4004. 48913 +0 0.046 4005. 48913 +0 0.046 400

HP(req’d) = [(37500)(20)] /[(6362)(0.75)] = 157 HP

Q P(gage) Density T(°F)1. 37500 -1 0.060 2002. 37500 -1 0.060 2003. 48913 -0.8 0.046 4004. 50562 -14.8 0.0445 4005. 48913 0 0.046 400

HP(req’d) = [(50562)(14.8)] /[(6362)(0.75)] = 157 HP

NOTE: A higher-volume, lower-pressure fan is required.

Fan Performance Requirements

Fan Inquiry Form

Attach gas composition and/or molecular weight if available.

Fan Type: ■■ Radial Blade ■■ Radial Tip ■■ Other ■■ BC ■■ AF ■■ FC ■■ OptionalSpecify ________________________________________________________________________________

Arrangement No. ___________________________________ Optional__________________________________________Rotation __________________________________________ Discharge ________________________________________

Inlets: ■■ Singles ■■ Double ■■ Optional Desired Noise Limit ________ dBA @_______________Ft.Inlet Boxes: ■■ Yes ■■ No ■■ OptionalDrive: ■■ Direct Coupled ■■ V-Belt Drive ■■ VFD ■■ Other __________________________________________Motor By: ■■ R I ■■ Others

Maximum Motor HP_____________________Preferred Speed __________Type _________Volts _______Phase ____________Cycle_____________________________

Accessories Yes (Y), No (N) or Optional (O)■■ Access Door ■■ Roller/Ball Bearings Special Requirements■■ Drain ■■ Casing ■■ Inlet Box ■■ Sleeveoil Bearings ■■ Air Perf. Test■■ Inlet Screen ■■ Split Housing ■■ Mech. Run Test■■ Flanged Inlet ■■ Blade Wear Protection ■■ Overspeed Test■■ Flanged Outlet ■■ Scroll Liners ■■ Sound Test■■ Drive Guard ■■ Side Liners ■■ Certified Matl. Test Reports■■ Heat Flinger ■■ Silencer ■■ Certified Welding■■ Shaft Seal Type __________ ■■ Circulating Oil System ■■ API 673■■ Insulation Clips ■■ Turning Gear ■■ Special Coatings■■ Insulated Housing ■■ Spray Nozzles Wheel ___________________■■ Radial Inlet Damper ■■ Special Paint, Coating Casing___________________■■ Louvered Inlet Damper ■■ Spring Isolation Base ■■ Spark Resistant■■ Outlet Damper ■■ Bearing Temp. Detectors AMCA-A, B or C _____■■ Independent Bearing Pedestals ■■ Bearing Vibration Detectors ■■ Pressure Test■■ Pedestal Sole Plates■■ Separate Damper, Size ______________________________________________________________________________

Special Requirements, Comments or Sketch: Attach Separate Sheet

Rep. Office ________________________________________ By ______________________ Date ____________________

Customer _________________________________________ Date Quotation Req.___________________________________Street ________________________________________City __________________________________________ State ____________________ ZIP _____________________Telephone _____________________________________ Fax ________________________________________________Attention ______________________________________

Reference or Project ________________________________Number of Fans Required ____________________________Description of Service ______________________________________________________________________________________

________________________________________________________________________________________________________________________________________________________________________________________________________

Max. Mechanical Design Temp ________ (Deg. F)

■■ Clean Air ■■ Dust/Moisture-Laden ■■ Abrasive ■■ Severe Abrasive ■■ Corrosive

Materials of Construction: Wheel_______________________ Casing______________________________________________Type of Dust or Gas ________________________________. _______________________________________________#/HRAltitude _______________________________Ft. Above S.L. Ambient Temp. Range _________to _________F

CONDITION INLET VOLUMEACFM OR SCFM

FAN SP,IN. WG

INLET SP,IN. WG

INLET TEMP.,DEGREES F.

INLET DENSITYLB/FT3


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Application Information

Radial Blade• Static efficiency to 75%• High tip speed capabilities• Reasonable running

clearance• Best for erosive or sticky

particulate

Backward Inclined• Static efficiency to 80%• Low to medium tip speed

capabilities

Radial Tip• Static efficiency to 75%• Medium to high tip speed

capabilities• Running clearance tighter

than radial blade but notas critical as backwardinclined and airfoil

• Good for high particulateairstream

Airfoil• Static efficiency to 87%• Medium to high tip speed

capabilities• Relatively tight running

clearances

Backward Curved• Medium to high tip speed

capabilities• High efficiency to 83%• Clean or dirty airstreams• Solid one-piece blade design

Paddlewheel• Open design/no shroud• 60-65% static efficiency• Inexpensive design• Good for high temperature

or highly erosiveapplications

• Medium to high pressure

Forward Curved(Sirrocco)• Smallest diameter wheel

for a given pressurerequirement

• High volume capability• 55-65% static efficiency• Often used for high

temperatures

Axial Flow• High volume, low pressure• 35-50% static efficiency• High temperature furnace

recirc. applications• Reversing flow capability• Airflow parallel to shaft axis


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Application Information


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Selection Information

Development of Fan Performance CurveReprinted from Publication 210 with the express written permission of the Air Movement and Control Association, Inc., 30 W. University Drive, Arlington Heights, IL 60004-1893.

A

SP

A

BC

D

E

E

D

CB

F

F

BHP

SP

CFM

Test duct orifice plates from shut-off to wide open for points A, B, C etc.

Typical Outlet Duct Test Setup

Notes

1. Dotted lines on fan inlet indicate an inlet bell and one equivalent duct diameter which may be used for inlet duct simulation. The duct friction shall not be considered.

2. Dotted lines on the outlet indicate a diffuser cone which may be used toapproach more nearly free delivery.


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Single Point OperationThe actual operating point will be determined by the intersection of the system resistance curve and the fan

performance curve. The fan must be selected correctly to exactly meet the design requirement.

System resistance curve

Fan performance curve

SP = 40"

BHP = 1257

Es = 0.75 (75%)

Es

BHP

THOUSANDS CFM

SP

STA

TIC

PR

ES

SU

RE

-IN

CH

ES

H2O

BH

PS

TAT

IC E

FF

ICIE

NC

Y, %

50

40

30

20

10

2000

1000

80

40

50 100 150 200 250

BHP = CFM x SP6362 x Es

(neglecting compressibility)

Fan Selection Considerations• Efficiency• Stability• Sound• Size• Speed


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ACFM x SP x Kp

6362 x static efficiency

ACFM = actual cubic feet per minute

SP = static pressure

Kp = compressibility factor

6362 = conversion constant

Static efficiency = , or

Fan HP =

ACFM x SP x Kp

6362 x Fan HP

Fan HP(input)

ACFM x SP x K p

6362

=Air H

P

(output)


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Regulation by Radial Inlet Damperor Inlet Box Damper (Parallel Damper)

Inlet damper control offers several advantages compared to outlet dampering. Because the air is prespun in the same angular direction as the fanwheel rotation, the energy required to operate the fan is significantlyreduced. Also, the multiple vanes just upstream of the fanwheel inlet provide a controlled presentation of air to thefanwheel that provides smooth control over a wide range of operation. With this system, there is, in fact, a new fanperformance curve for every damper position. See the example below:


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Example: (Fans with geometric similarity)

100" diameter wheel x 10" tip width100,000 CFM @ 20" SP @ 70° F @ 393 BHP

What is performance at 90-1/2" diameter wheel x 9.05" tip width?

CFM2 = (90.5/100)3 (100,000) = 74,122 CFM

SP2 = (90.5/100)2 (20) = 16.3 in. H2O

BHP2 = (90.5/100)5 (393) = 239 HP

Volume: CFM2 = (Size2 / Size1)3 (CFM1)

Static Pressure: SP2 = (Size2 / Size1)2 (SP1)

Horsepower: BHP2 = (Size2 / Size1)5 (BHP1)

Sound Power: LW2 = LW1 + 70 log (Size2 / Size1)

Size Change Fan Law

The Size Change Fan Laws make it easy to determine the performance for a larger size fan based on the knownperformance of an existing fan or a laboratory model fan. This requires that the fans be geometrically similar. Notehowever that Robinson has developed modified versions of this fan law that allow accurate prediction of tipped-out(slightly larger diameter) and de-tipped (slightly smaller diameter) fan rotors. This is sometimes a very cost-effectivemeans of accomplishing in-field modifications to increase performance or decrease horsepower consumption.

Example: (Fans to be tipped out or de-tipped)

CFMMod = (Dia2/Dia1)2 (CFMorig)

SPMod = (Dia2/Dia1)2 (SPorig)

BHPMod = (Dia2/Dia1)4 (BHPorig)

Example: 100" diameter wheel x 10" tip width100,000 CFM @ 20" SP @ 70° F @ 393 BHP

What is performance with a 5% tipout to 105'' diameter?

CFMMod = (105/100)2 (100,000) = 110,250 CFM

SPMod = (105/100)2 (20) = 22.05 in. H2O

BHPMod = (105/100)4 (393) = 478 HP

Note: Effect on sound pressure is greater due to decreased cut-off clearance.


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Erosion

Erosion is a major problem on some induced draft fans. High dust loading combined with high inlet velocities canresult in dangerous shutdowns. Hard surface wear liners can help. But reducing the fan inlet velocity by specifying alower speed, larger diameter fan can result in reduced wear and extended fan wheel life.

The kinetic energy (KE) of the particles decreases with the square of the velocity.

Erosion can also be reduced by reducing the particulate flow rate (lbs./hour) or by reducing the averageparticle size.

1. Reducing the particulate flow rate (lbs./hour).

2. Reducing the average particle size.

3. Reducing the particle velocity and the gas velocity at the fan inlet.

Note that changing from a single inlet fan design to a double inlet design is a very effective means of decreasing thefan inlet velocity.

Note: The presence of particulate matter in the gas stream affects the average density at the fan inlet and therefore themotor power requirement as is shown on page 1.


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(Note: Over 180 different materials tested to date)


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Comparative Advantages of Oil and Grease

Advantages of Grease1. Maintenance work is ordinarily reduced since thereare no oil levels to maintain.2. Grease in proper quantity is more easily confined to the housing. Design of enclosures can therefore be simplified.3. Freedom from leakage is readily accomplished infood, textile, chemical industries and where contami-nation of products must be avoided.4. Grease improves the efficiency of labyrinth enclo-sures and offers a better protection for the bearing.5. Bearing can be installed in a high velocity gas stream.

Advantages of Oil1. Oil is easier to drain and refill. This may be more desirable for applications requiring shortlubricating intervals.2. The correct amount of lubricant is more easily controlled.3. Oil lends itself more readily to the lubrication of allparts of a machine.4. Oil lends itself to applications with highertemperatures.5. The bearing friction and temperature rise are usuallymore favorable.


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Sleeve-Type Bearings• No moving parts• No metal-to-metal contact/infinite fatigue life• Babbitted surfaces• Split design for easy maintenance• Shaft-mounted thrust collar (fixed bearings)• Infinite axial expansion capability (for shaft)• External cooling (water, air, circulating oil)• Auxiliary dust seals available

Note: These bearings are generally used for large diameter fans. They have higher load and speed capabilitiesthan rolling element bearings.

Fan Bearing Comparison

Standard SphericalItem of Ball Bearing Roller Bearing Dodge Sleeve

Comparison (Deep Groove) (SAF Housing) Bearing

1. Calculated Life 40,000 / 80,000hrs. 80,000 hrs. Infinite2. Radial Load Capability Low Medium Very Large3. Axial Load Capability Low Medium Very Large4. Cooling Capability Limited Limited Very High

water cooling No No Yesair cooling No No Yes

5. Circulating Oil No Yes Yesmax. flow rate N/A Limited Very High

6. Static Oil Lube No Yes Yes (5-7/16" brg. = (5-7/16" brg.

.6 liters) XC = 2.95 L / RT = 4.49 L)

ring oiled No No Yes7. Grease Lubrication Yes Yes No8. Thermal Analysis Yes Yes Yes9. Dynamic Stiffness Medium High Medium (oil film)

10. Dynamic Damping Minimal Minimal Good11. Shaft Tolerances Nominal Nominal Nominal

+0.000 / -0.0005 +0.000 / -0.005 +0.000 / -0.00212. Shaft Surface 32 32 3213. Shaft Specialties None None XC – None

RT – groove/collar14. Self-Aligning Yes Yes (brg.) Yes (liner to housing)15. Axial Expansion 3/16" – 3/8" Std = 3/8" Unlimited16. Split Housing No Yes Yes17. Split Bearing No No Yes18. Field Replacement Easy Easy Easy

(Outboard)19. Field Replacement Difficult Difficult Fairly Easy

(Inboard) (Must remove coupling (Must remove coupling (Loosen coupling, jackand move motor before and move motor before up shaft about 1/4",

removing/installing bearing) removing/installing bearing) slide old bearing out)20. Spare Parts Worldwide Worldwide USA only21. Availability Stock Stock Stock22. Axial Length Shortest Medium Longest

(affects shaft sizing)23. Price Low Low High


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D = shaft dia. (in.)I = πD4

(in.4)64

E = Shaft modulus of elasticity (lbs./in.2)W = fanwheel weight (lbs.)A, B, L = distance (in.)y = shaft deflection (in.)Ncr = critical speed (rpm)

y = WL3

48EI


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Fanwheel DesignFatigue LifeASME fatigue life curves are used for both high-cycle and low-cycle fatigue calculations. The following is an examplefor ASTM A514 material.

Ab

sorb

ed E

ner

gy

ft.-

lbs.

Str

ess

(psi

)

Impact ToughnessCrack propagation rate is a function of the toughness of the material. This can be quantified by Charpy V-Notchtesting. Materials with low impact toughness are more glasslike or brittle.

The impact toughness varies as a function of temperature for many materials. There is a history of rapid, unexpected failures of fan rotors during cold weather start-up or operation that has been related to low materialimpact toughness. Therefore, Robinson recommends using materials that have an adequate impact toughness at the lowest expected operating temperature. The graph below shows that thermal stress relief of welded A514material has a detrimental effect on the impact toughness. Therefore, thermal stress relief of A514 fanwheels is not a recommended practice. Other materials, of course, may not react to thermal stress relief in this way.

107

106

105

104

10 102 103 104 105 106


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Shaft Seals:These are designed to minimize leakage of dangerous gases from the process, or to exclude atmospheric air fromthe process itself. Several seal designs are available with varying sealing effectiveness. (See section on shaft sealdetails).


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Shaft may be chrome-plated or ceramic-coated

under packing area.


under packing area.


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under carbon.


under carbon.


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In addition to the sound level, the duration of exposure toa particular noise level is also an important local noiseconcern. The table below outlines OSHA’s permissiblenoise exposure limits.

Duration per day, hours Sound level dBA

8 ..................................................................906 ..................................................................924 ..................................................................953 ..................................................................972 ................................................................1001-1/2 ..........................................................1021 ................................................................1051/2 ..............................................................1101/4 or less ..................................................115

When the daily noise exposure is composed of two or more periods ofnoise amounts of different levels, their combined affect should beconsidered, rather than the individual effect of each. If the sum of thefollowing fractions — C1/T1 + C2/T2… Cn/Tn — exceeds unity, then themixed exposure should be considered to exceed the limit value. Cnindicates the total time of exposure at a specified noise level, and Tnindicates the total time of exposure permitted at that level.

Federal Register, Vol. 34, No. 96, May 20, 1959, pp. 7849.

For a ducted inlet/ducted outlet fan in a free field:

LpTOTAL = 10 log (10 + 10 + 10 + 10 )Lpd

10

Lpe

10Lpf

10Lpm

10

LpTOTAL = 10 log (10 + 10 + 10 + 10 )8310

8810

8110

8510

For example:

LpTOTAL = 91.04dB


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time (seconds)delta RPM = change in speed (rev./min.)available torque = (motor torque capability) - (fan torque

requirement) at all speeds from zero to normal operating speed (ft.-lb.)

WR2 = fan rotor rotational moment of inertia (lb.-ft.2)


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Fabrication & Quality Control


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Notes:1. The numerical value after the letter G is equal

to the product of ePER(2πf) expressed in millimeters per second.

G = ePER (2πf), or ePER =

UPER = max residual unbalance for a particular rotor.

UPER = ePER x m = Gm2πf

Example: Rotor weight, (m) = 1000 lbs; Rotor Dia = 75 inch;Maximum operating frequency = 1200 rpm ÷ 60 s/min = 20 cyc/s;Balance to grade G2.5 = 2.5mm/sec.

ePER = = = 0.02 mm

0.02 mm ÷ 25.4 mm/in = .0008 in

UPER = ePER x m = (0.0008 in)(1000 lb) = 0.8 lb-in

0.8 lb in x 16 oz/lb = 12.8 oz-in

Shaft* vibration in free space =

G2πf

G2π (20)

( )2.5mm/sec25.4mm/in = 0.10

insec

*Note: Bearing vibration will be considerably lower thanshaft vibration.

G(2πf)

(peak velocity)


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Safety Relief

Safety Relief

PlantAir/H2O

PlantAir/H2O

Pressure Test Procedure for Fan Casings(When no leakage rate is included in the specifications)1. Blind flange with rubber gasket over all inlet and outlet openings.2. Fans with a peak operating static pressure from 0 in-H2O to 138 in-H2O to be pressure tested with compressed

air. Fans with peak operating static pressures over 138 in-H2O to be pressure tested with water.3. Attach plant air line or water line to casing drain or other entry point as follows:

4. Use U-tube water gauge or a 0-5 psi pressure gauge for pressure of 0 to 50 in-H2O. Use 0-5 psi pressure gauge for pressures above 50 in-H2O.

5. Pressurize casing to highest operating static pressure of the fan. (Examine all operating conditions, including 70° F, and use the highest pressure.)

6. Close inlet valve.7. All welds are to be bubble tested with leak detection solution. Repair all weld leaks and begin retesting.8. On the Pressure Test Log Sheet (See page 90), record pressure once every 60 seconds for 15 minutes.9. If average leak rate is less than 2.0 in-H2O per minute, then the unit is satisfactory. If the average leak rate

exceeds 2.0 in-H2O per minute, inspect the area of the shaft seal, inlet/outlet flanges and casing split gaskets first for major leaks.

10. Continue to repair leaks and rerun pressure test until the required leak rate is achieved.


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Pressure Test Log Sheet

Factory Order # Serial # Test Date

Size and Description

Assembly Drawing #

Shaft Seal Material Gasket Material

Highest Operating Static Pressure of Fan (all operating conditions, including 70° F) in.-H2O

Test Pressure (max.) in.-H2O

Casing Welded: Inside ■■ Outside ■■ Both Sides ■■


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Installation Information


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Troubleshooting


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Troubleshooting


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Troubleshooting


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Troubleshooting


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Troubleshooting


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Troubleshooting


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Engineering Publications

Engineering Publications1. Gutzwiller, H.L., Robinson Industries, Inc. “Balance

and Vibration Considerations for Fans,” AMCATechnical Seminar, Los Angeles, California,November 1-3, 1995.

Abstract: For years, we have reviewedspecifications from users of industrial fans thatshow a misunderstanding between the terms“balance” and “vibration.” While the two termsare related, they have independent definitions;and it is important for both users andmanufacturers to clearly understand thedifferences. The Air Moving and ControlAssociation (AMCA) has formed a committee todraft a standard #204 entitled “Balance Qualityand Vibration Levels for Fans.” Other standardsexist for defining balance and vibrationrequirements for the general rotating equipment.However, this committee is to address thespecial consideration for fan equipment inparticular. The purpose of the standard is to“define appropriate fan balance quality andoperating vibration levels to those who specify,manufacture, use and maintain fan equipment.”

2. Banyay, H.D., Robinson Industries, Inc. “MaximizingFan Reliability in Kilns, Dryers and Burners,”Ceramic Industry, Vol. 146, No. 4, April 1996,pages 5-15.

Abstract: The purpose of this paper is todiscuss the proper application of bearings on hotgas fans. The gas temperature for these canrange up to 1300° F or more. The fans typicallyare V-belt driven (AMCA arrangement #1 or #9)and range up to 100 HP.

3. Gutzwiller, H.L., and Banyay, H.D., RobinsonIndustries, Inc., and Ball, W.D., John ZinkCompany. “Marine Vapour Recovery SystemBlowers,” Fans for Hazardous Applications,Seminar by the Fluid Machinery Committee ofthe Institution of the Mechanical Engineers,London, England, October 4, 1994.

Abstract: This paper describes the design andapplication of high-pressure blowers used toexhaust combustible hydrocarbon vapors fromseagoing tankers. Important considerationsinclude stable operation over a wide flow range,spark-resistant and gas-tight construction,minimal noise, resistance to salt-spray corrosion,and ease of maintenance.

4. Grupp, David, Robinson Industries, Inc. “NaturalFrequency of a High Temperature Plug Unit andWall in a Furnace Application,” EngineeringPaper 2566-94-A1, AMCA EngineeringConference, St. Petersburg Beach, Florida,February 20-22, 1994.

Abstract: Critical speed is a well-known designparameter in the fan industry. Most commonly,critical speed is related only to the fan rotor andshaft assembly. Often the effects of the bearings,support structure, foundation and soil areneglected as properties of the system. In mostcases, the stiffness of each of these properties isso high that their effect is indeed negligible.However, when the effects of these propertiesbecome significant, the fan engineer must becareful to design for the system critical speed.The following paper will present a fan applicationproblem in which the stiffness of a wall in afurnace structure resulted in a unique systemcritical resonance at the operating speed of fivehigh temperature axial flow fan assemblies.

5. Gutzwiller, H.L., and Banyay, H.D., RobinsonIndustries, Inc., and Cohen, S.N., FullerCompany. “Cement Plant Preheater Fan Build-up Control,” IEEE Conference, Tarpon Springs,Florida, May 22, 1990.

Abstract: In recent years, greater demands withregard to throughput and efficient utilization ofheat in the kiln due to process design changeshave placed greater demands on the kiln-induced draft fan. These fans have beendesigned with ever-increasing volume and staticpressure requirements, as well as higherprocess gas temperatures. This, of course,means larger fan rotors operating at very high tipspeeds. Along with these design changes, theproblem of build-up on the impeller has alsoincreased markedly … But why must someplants battle this problem routinely while othershave no significant ID fan build-up at all? Whatcauses kiln ID fan build-up problems? How can itbe stopped?


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Documents

Engineering Manual.pdf