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.

Lets 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. Its 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, letsconsider 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

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


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. n

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)(gpmft) x 0.15 = Ns(m3/minm)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


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 CURVE

The system curve represents thepressures 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 directly 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 dont 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


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. n

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.


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 throttlevalvessituations 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 users 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

800600500400300

200

10080605040

30

20

108

6543

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

2014

1052 05

985

800

600500

400300

200

140

100

-60 to 240F

TEMPERATUREF

CARBON

DIOXID

E

NITRO

US OXI

DE

ETHAN

E

MONO

CHLOR

OTRIFLU

OROM

ETHANE

HYDR

OGEN

SULFI

DE

PROPYL

ENE

PROP

ANE

AMMO

NIA

CHLORI

NE

METH

YL CH

LORID

ESU

LFUR D

IOXIDE

ISOBU

TANE B

UTAN

E

ETHY

L CHL

ORIDE

METH

YL FO

RMAT

E

DIETH

YL ET

HER

METH

YLEN

E CHL

ORIDE

DICH

LORO

ETHY

LENE

ACET

ONE

DICH

LORO

ETHY

LENE

(CIS)

CHLO

ROFO

RM (TR

ICHLOR

OMETH

ANE)

CARB

ON TE

TRAC

HLOR

IDE TRIC

HLOR

OETH

ULEN

E

WATE

R

HEAV

YWAT

ER (SP

.GR. AT

70 F=

1.106

GAU

GE

PRES

SURE

LBS

. PER

SQ.

IN.

VACU

UMI

NCHE

S O

F M

ERCU

RY

ABSO

LUTE

PRE

SSUR

ELB

S. P

ER S

Q. IN

.


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 2030% 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

POIN

T O

FLO

WES

T PR

ESSU

REW

HER

E VA

PORI

ZATI

ON

STAR

TS

INCR

EASE

PRE

SSUR

E

4a SUCTION SUPPLY OPEN TO ATMOSPHERE-with Suction Lift

CL

PB NPSHA=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


(23 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 designersNPSHA. 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%

NPS

HR

TOTA

L HE

AD

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


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. n

REFERENCES1. Goulds Pump Manual.

2. Durco Pump Engineering Manual.

3. Hydraulic Institute TestStandards1988 CentrifugalPumps 1.6.

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

HEA

D

NPS

H - F

EET

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 16

18 2028

32 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


f a wide receiver has the rightspeed and good hands, all thatsneeded 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, allthats 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 quarterbacks pass is offtarget, badly timed, or the ballsturning end over end in the air,the receiver may not be able tohang on to it, and theres no gainon the play. When that hap-pens, the quarterbackknows he didnt throw itproperly and doesnt blamethe receiver. Unfortunately,thats where the compari-son ends. The engineeringquarterbacks tend toblame the pump even whenits their delivery thats 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,lets 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. Lets simplify it.There isnt 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

HfHs


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 cant 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


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 satisfactorypump operation is obtained, that

doesnt 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 arent performing well,is it their fault? Or does the quar-terback need more training? n

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


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 D T in F isexpressed as:

H 1D T = x

778 x Cp h 1

where

H = total head in feet

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

h = pump efficiency in decimalform

778 ftlbs = energy to raise thetemperature of one pound ofwater 1F

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 15F. 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 D 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


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 manufacturersperformance 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 humprequires an impeller redesign; how-ever, note that there will be a loss ofefficiency and an increase in NPSHR.

Whats 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 shuttlesbetween 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.


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, dont 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 dont 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. Frasers 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 areab 1 = impeller inlet angleb 2 = 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 Calculationsfor details

Cm2U2

Discharge Angle b 2 Inlet Angle b 1

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


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). n

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-lors 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.


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) H3H2

H1

System-Head Curve

FrictionLossesStaticPressur eor Head

}Head

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


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

EXAMPLE

Lets 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

CH-Q 1800 R.P.M.

143/4"Impeller

Q1

43 /4"

Impe

ller

Q14

3 /4"Impel

ler

14"Impeller

Q1

4"Im

pelle

r

Q14"

Impelle

r

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

fficie

ncy

Feet

Tot

al H

ead

FIGURE 4


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. n

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 CBHP 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 AImpeller 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)


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 fluids 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 casescertain crude oils,for examplethis is not the case.

Heavy crude oils can haveviscosities high enough to increasethe friction drag on a pumpsimpellers significantly. The addi-tional horsepower required toovercome this drag reduces thepumps 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:

m = AeB/T

where

m = 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


perature drop of 1C.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 Americas 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 )


The process of determiningthe effect of a fluids 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 = KD2g u3

where

K = a constant based on the Reynoldsnumber

D = impeller diameter

g = fluid density

u = impeller tip speed

Like most of Stepanoffs 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 pumps 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 Dx6C

h

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

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/CQThe next step would be to

find a pump having the requiredperformance on water. Afterdetermining the efficiency of thepump on water, we would correctit for the viscous case as shownabove:

h viscous service = h water x ChThe horsepower required by

the pump at this point would becalculated as follows:

hpviscous service =

(Qviscous service x Hviscous service x sp gr)

(3,960 x h viscous service)

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

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.


Ch

CQ CH0.6 CH0.8 CH1.0 CH1.2Per 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


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 therotors 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 rotors 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 e Wj 4 Wjwhere: N2 NWj = rotor weight per balance plane, Ibf N = rpme = eccentricity, in.Eccentricity (e ) or Specific Unbalancein.oz/lbm 56347 16 e 4

N2 N

in.lbm/lbm 3522 e 0.25N2 N

where RUB = e Wj see Table 2Unbalance Force (UBF),lbf where:UBF = e Mw 2 0.10 Wj e WjN2 WjNand M = Wj/386 lbfs2/in. 35200 140800w = 2 p N/60 rad/sCircular Velocity (CV),in./s 368 e N 0.26

N 9.54

mm/s 9347 2.66 e N 0.665N

where CV = e w ISO Standard 1940 G 9347 G 2.66 e N 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 1940Balance Quality ofRigid Rotors. It also forms thebasis of API Standard 610 7thEditions 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 rotors center of mass toits center of rotation. Note that 1in.oz is equivalent to 72.1 cmg.

ECCENTRICITYThis is the distance that the

center of mass of the rotor is dis-placed from the rotors 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 oflengthmass/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


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 Editions 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 rigidrotorsthose 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 Standard5406The Mechanical Balancing ofFlexible Rotors and ISO Standard5343Criteria 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 e , m -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 rotors 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 G6.3 and G2.5 formost pump components, whereAPI 610 calls for the equivalent ofG0.67. 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, e , to account for theeffect of system dynamics.Algebraically, the physics of thesituation can be represented asfollows:

x = X sin (w t F )

(w /w n)2X = e

([1 (w /w n)2]2 + (2zw /w n)2)0.5

2zw /w nF = tan1

1 (w /w n)2wherex is the displacement of a point on

the rotorX is the magnitude of the vibration

at that pointe is the mass eccentricityw is the operating speed or fre-

quency of the rotorF is the phase angle by which the

response lags the forcez is the damping factor for the

mode of vibration under consid-eration

X/e is the amplification factorw /w n 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. n

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

TABLE 2. BALANCE QUALITY CLASSES


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 rotorloadsand 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


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 oclock.

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. F

Centrifugal Pumps Handbook

Documents

Transcript of Centrifugal Pumps Handbook