Libro de Bombas

7/29/2019 Libro de Bombas

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

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

VAPOR PRESSURE

Knowledge of vapor pressurei s ext r emely impor tant whenselec t ing pumps and the i rmechanical seals. Vapor pressure

is 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.

CAVITATION

Cavitation can create havoc w ithpumps 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 pum p,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 suction

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

The NPSH Amust be equal to orgrea ter than the

where PB = barometric pres-sure in feet absolute, VP = vaporpressure of the liquid at maximumpumping t empera ture in f ee t

absolute, 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 barometr icpressure), and hv = velocity headin the suct ion pipe in feetabsolute.

NPSH R can only be deter -mined during pump testing. To

determine i t , 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 on

the test floor. Its during thesetests that one gains a r eal 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:

H e a d i s a t e rm used toexpress pressure in both pumpdesign and system design whenanalyzing static or dynamic condi-t ions . This relat ionship isexpressed as:

h d i f t(pressure in psi x 2.31)

Nomenclature and DefinitionsBY PAT FLACH

V

FIGURE 1

C EN TR I F U GA L P U M P S

HANDBOOK

100 100 100


2/260

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 measur ingpounds per square inch at thebottom of the column, not the

total 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 he ad 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 suction

nozzle 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 n ozzle down to the sur faceof the suction supply liquid.

Velocity, friction, and pressurehead are used in conjunction with

static 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 same

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 determ ined in the fieldby taking the discharge pressurereading, converting it to head, andcorrecting it to the pump center-l ine, then adding the veloci ty

head.

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 suction

resistance can come from pipe fric-tion, valves, and fittings. Values infeet of liquid can be found in theHydraulic Insti tute 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 1.13

specific gravity

When discussing how a pumpperforms in service, we use termsdescribing dynamic head. In otherwords when a pump is running it is

FIGURE 2

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

TotalStaticHead

StaticDischarge

Head

StaticSuctionHead

StaticDischargeHead

TotalStaticHead

StaticSuction

Lift


3/260

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 (total

dynamic head) as the liquid passesthrough the casing or diffuser.

To predict the approximate totaldynamic head of any centr i fugalpump, 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= revolut ions 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 v

from 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 diameter

will 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. s

Patrick M. Flach is the western

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, m ore than English units, it still presents

a problem to a lot of engineers.Authors are en couraged 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 conver sions frequen tlyfound in solving hydraulic engineering problems as they relate topumping systems.

PUMPING UNITS

FLOW RATE

(U.S.) gallons/min (gpm ) x 3.785 = l iters/min (L/min)

(U.S.) gpm x 0.003785 = cubic meters/min (m 3/m in)cubic feet/sec (cfs) x 0.028 = cubic m eters/sec (m 3/s)

HEAD

feet (ft) x 0.3048 = meters (m)pounds/ square inch (psi) x 6,89 5 = Pascals (Pa)

POWER

horsepower (Hp) x 0.746 = kilowatts (kW)

GRAVITATIONAL CONSTANT (g)

32.2 ft./s2 x 0.3048 = 9.81 m eters/second2 (m/s2)SPECIFIC WEIGHT

lb/ft3 x 16.02 = kilogram/cubic meter (kg/m 3)

VELOCITY (V)

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

VELOCITY HEAD

V2/2g (f t) x 0 .30 48 = m eters ( 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 SystemsEditorial Advisory Board.

Pumping Terms


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Centrifugal and Positive Displacement Pumpsin the Operating System

n the many differences that existbetween centrifugal and positive

displacement pumps, one whichhas caused some confusion is themanner in which they each operatewithin the system.

Positive displacement pumps havea series of working cycles, each ofwhich encloses a certain volume off luid an d m oves i t 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 bas ic des ign 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 w ith the cen-t r i fugal pump, which can onlyreact to the pressure demand of the

system. 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 of

can develop is reduced as the capacityincreases. Conversely, as the capacity

drops, 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 liquidsof 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 flow

h h ld h i h

When the pump curve is super-imposed on the system curve, the

point 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 , i tshould be noted that there are onlya few conditions which will cause

the pump curve to change its posi-tion and shape:

wear of the impeller change in rotational speed change in impeller diameter change in liquid viscositySince these conditions dont nor-

mally develop quickly, any suddenchange in pumping conditions islikely to be a result of a movement

in the system curve, which meanssomething in the sys tem 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 any

I

BY ROSS C. MACKAY

Pump Curve

System Curve

H

OQ


HANDBOOK


5/260

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.

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. s

Ross C. Mackay is an independent con-

sultant located in Tottenham, Ontario,Canada. He is the author of several papers

on the practical aspects of pump mainte-

nance, and a specialist in helping companies

reduce their pump maintenance costs.


6/260

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

Bubble format ion

occurs at a point where the pres-sure is less than the vapor pres-sure , and bubble col l apse 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 in

other hydraulic systems such asbypass or i f ices and throt t levalvessituat ions where anincrease in velocity with resultingdecrease in pressure can reducepressure below the liquid vaporpressure.

CAVITATION EFFECTS

BUBBLE FORM ATION PHASE

Flow is reduced as the liquidis displaced by vapor , andmechanical imbalance occurs asthe impeller passages are partiallyfilled with lighter vapors. Thisresul ts in vibrat ion and shaf tdeflection, 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 removesegmen ts of impeller material.

2 N i d ib ti lt f

Cavitation and NPSH in Centrifugal PumpsBY PAUL T. LAHR

CFIGURE 1

involves both the net positive suctionheads avai lable in the sys tem(NPSHA) and the net positive suctionhead required by the pump (NPSHR).

NPSHA is the measurement orcalculation of the absolute pressureabove the vapor pres sure a t 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.


HANDBOOK

1000

800

600

500

400

300

200

100

80

60

50

40

30

20

10

8

6

5

4

3

2

1.0

.80

60

28"

28.5"

10"

15"

20"

22.5"

25"

26"

27"

80

60

50

40

30

20

14

1052

05

985

800

600

500

400

300

200

140

100

-60 to 240F

CARB

ONDIOXID

E

NITR

OUSO

XIDE

ETHA

NE

MONO

CHLO

ROTR

IFLUO

ROME

THANE

HYDR

OGEN

SULFID

E

PROPYLEN

E

PROP

ANE

AMMON

IA

CHLO

RINE

METHYL

CHL

ORIDE

SULFURDIO

XIDE

ISOBU

TANE BU

TANE

ETHY

LCHL

ORIDE

MET

HYLF

ORMATE

DIETHY

LETH

ER

METH

YLEN

ECHLO

RIDE

DICHLO

ROETHYLEN

E

ACETON

E

DICH

LORO

ETHY

LENE

(CIS)

CHL

OROF

ORM(TR

ICHLO

ROME

THAN

E)

CARB

ONTETR

ACHLO

RIDE TR

ICHLO

ROETHU

LENE

WATER

HEAVYW

ATER

(SP.GR

.AT70

F=1.1

06

GAUG

EPRESSURELBS.PERSQ.IN.

ERCURY

ABSOLUTEP

RESSURELBS.PERSQ.IN.


7/260

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 turbu lence.

Turbulence loss is extremelyhigh a t low f low and thendecreases with flow to the best

efficiency 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 NPSH R for thepump at the stated flow.

The pum p m anuf ac t u r e rdetermines the actual NPSHR foreach pump over i t s completeoperating range by a series oftests. The detailed test procedureis descr ibed in the Hyd ra ulic

Inst it ute Tes t St an da rd 19 88Centrifugal Pumps 1.6. Industryhas agreed on a 3% head reduc-tion at constant flow as the stan-dard value to establish NPSH R.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 the

FIGURE 2

FIGURE 3

Calculation of system net positive suction head available (NPSH A) for typicalsuction conditions. PB = barometric pressure in feet absolute, VP = vaporpressure of the liquid at maximum pum ping 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

ENTRANCELOSS

FRICTION

TURBULANCEFRICTION

ENTRANCELOSS AT

VANE TIPS

INCREASINGPRESSURE

DUE TOIMPELLER

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


8/260

(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 NPSH A,and normal operating flow mustbe considered.

SUCTION SPECIFIC SPEED

The concept of suction specif-ic speed (Ss) must be consideredb y th e p u mp d es ig n er , p u mpapplication engineer, and the sys-tem designer to ensure a cavita-t ion-f ree pump wi th h ighreliability and the ability to oper-ate over a wide flow ran ge.

N x Q0.5

Ss = (NPSHR)0.75

w here N = pum p rpm

Q = flow rate in gpm at thebest efficiency point

NPSHR = NPSHR at Q withthe maximum im pellerdiameter

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, designers

and 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 NPSH Aor NPSH R at 3,500 and1,750 rpm. Similar plotscan be made for o ther 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 NPSH A, 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 allow

flows to 4,000 gpm with 20 feet ofNPSHA.

Q1

Q2

100% CAP Q3

Q4

NPSH

3%

NPSHR

TOTAL

HEAD

FIGURE 4

Typical results of a four-point net posi-

tive suction h ead required (NPSHR) testbased on a 3% head d rop.

3

2

19

8

7

6

5

4

HSV=

2

3

4

5

6

7

89

10

HSV=12

28

32 5055

HSV=

24

45ons

pecific

speed

FIGURE 5


9/260

It is important forthe pump user to under-stand how cri t ical thesystem design require-ments are to the selec-t ion of a rel iable,trouble-free pump.

Matching the system

and pump characteristicsis a must. Frequently,more attention is paid tothe discharge side. Yet itis well known that mostp u mp p er fo rman cep r o b l e m s a r e c a u s e db b l h

As a general rule, the higherthe suction specific speed, thehigher the minimum stable flowcapacity wil l 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 reduced

flow, then the Ss value must begiven careful consideration. s

REFERENCES

1. Goulds Pump Manual.

2. Du rco Pu mp En g in eer in gManual

HEAD

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

4

3

2

1

98

7

6

5

4

3

2

1

HSV=

12

HSV=

1

109

8

6

3

2

14

4

5

716

18 2028

32 36 40

50

HSV=

24

HSV=

45

Solution for

S=N

for N=1,750 rpm

Q

Hsv0.75

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.) H SV= NPSHR atBEP with the maximum impeller diameter.

Q, Capacity, gpm

S,

Suctions

pecific

speed

FIGURE 6


10/260

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 even

laminar flow into the eye of theimpeller.If the quarterbacks pass is off

target, 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 quar t erbackknows he didnt throw itproperly and doesnt blamethe receiver. Unfortunately,thats where the compari-son ends. The engineeringquar terbacks tend toblame the pump even whenits their delivery thats bad!

Just as there are tech-niques a quarterback mustl earn in order to throwaccurately, there are rules

which ensure that a liquidarrives at the impeller eye withthe pressure and flow characteris-tics needed for reliable opera tion.

RULE #1.PROVIDE SUFFICIENT NPSH

Without getting too complicat

a function of the system design onthe suct ion s ide 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 temptat ion 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 + H s - Hvp - Hf

whereHa= the head on the surface of the

liquid in the tank. In an opensystem like this it will be

Hf= t he fr ict ion losse s in t h esuction p iping.

The NPSH Available may alsobe determined with this equation:

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

where

Ha= a tm osp h er ic pr essu re infeet of head.

Hg= t he gauge pr essu re at thesuction flange in feet ofhead.

V2= Th e v elo cit y h ea d a t t hepoint 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 par t icular l iquid being

pumped. Many industries workwith a maximum velocity of about5ft./sec., but this is not alwaysacceptable.

RULE #3.NO ELBOWS ON THE

SUCTION FLANGE

Pump Suction ConditionsBY ROSS C. MACKAY

I

FIGURE 1

2g


HANDBOOK

Ha

Hvp

Hf

Hs


11/260

entrainment, which may result inimpeller damage and vibration.

When the elbow is installedin a horizontal plane on the inleto f a doub l e s uc t i on pum p ,uneven flows are introduced intothe oppos ing 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 a

pump 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-na r f l ow t o t he eye o f t heimpeller 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 ont he l a r ge r p i pediameters and viceversa.

RULE #4. STOP AIROR VAPOR ENTERINGTHE SUCTION LINE

Any high spotin the suction linecan become f i l ledwith air or vapor which, if trans-ported into the impeller, will createan effect similar to cavitation and

wi th th e same resul t s . Serviceswhich are particularly susceptibleto this situation are those where thepumpage conta ins a s igni f i cantamount of entrained air or vapor,as well as those operating on a suc-tion lift, where it can also cause thepum p to lose its prime. (Figure 3)

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

t ioned wi ththe f lat s ideu p p e r m o s t .(Figure 4).

If a pum pis t ak ing i t ssuction froma s u m p o r

tank, the for-mation of vortices candraw air into the suc-tion line. This can usu-a l ly be prevented bypr ov i d i ng s u f f ic i en tsubmergence of liquid

h i

tices are more difficult to trou-bleshoot in a closed tank simplybecause they cant be seen as

easily.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 PIP ING ALIGNM ENT

FIGURE 2

FIGURE 4

FIGURE 3

Air Pocket


12/260

Under certain conditions thepump manufacturer may identifysome maximum levels of forcesa n d m o m e n t s w h i c h m a y b eacceptable on the pum p flanges.

In high temperature applica-tions, some piping misalignment

i s inev i t ab l e ow i ng t o t he r m a lgrowth during the operating cycle.Under these conditions, thermalexpansion joints are often intro-duced to avoid transmitting pipingstrains to the pump. However, ift he en d o f t he expans ion j o in tc l o s e s t t o t h e p u m p i s n o tanchored securely, the object ofthe exercise is defeated as the pip-i n g s t r a i n s a r e s i m p l y p a s s e dthrough to the pump.

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

Piping designis one area wherethe bas ic pr inci -ples in-volved areregularly ignored,r esul t ing 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 fai l , the t rouble is rarelytraced to incorrect p iping.

I t has been a r gued t ha tbecause many pumps are pipedincorrectly and most of them areoperating qu ite 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 sa t i s f ac torypump operation is obtained, that

doesn t au tomat ica l ly make aquestionable piping practice cor-rect. It merely makes it lucky.

The suction side of a pump ismuch m ore impor tant than thepiping on the discharge. If anymis takes are made on the dis -charge side, they can usually becompensated for by increasing

the 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.

I n o t he r w or ds , i f you rreceivers arent performing well,is it their fault? Or does the quar-terback need more training? s

Ross C. Mack ay is an indepen-dent consultant who specializes inadvanced technology training for

pum p maintenan ce cost reduction.He also serves on the editorial adviso-ry board forPumps and Systems.

Inflow Inflow

To PumpSuction

To PumpSuction

Baffle

FIGURE 5


13/260

i n imum f low 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 (minimumtherm al flow)

2. Minimum stable flow

3. Thrust capacity

4. NPSH requirements

5. Recirculation

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

TEMPERATURE RISETemperature rise comes from

energy impar ted to the l iquidthrough hydraulic and mechanicallosses within the pump. Theselosses are conver ted to heat ,

which can be as sumed to beentirely absorbed by the liquidpumped. Based on this assump-tion, temperature rise T in F isexpressed as:

H 1T = x

778 x Cp 1

whereH = total head in feet

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

= pump efficiency in decimalform

mechan ical handb ooks.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 di f ferencebetween the allowable temperatureand the temperature at the pumpinlet is the maximum allowable

temperature rise. Knowing T andCp , the minimum f low 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 M inimum FlowBY TERRY M . WOLD

M

SUCTION

Low PressureLower

Temperature

DISCHARGE

High PressureHigherTemperature


HANDBOOK

FIGURE 1


14/260

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 vapor

pressure 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 (C

p= 1.0) that

can 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 pump

is built, options for retrofitting thepump with a heat shield are greatlyreduced.

M INIM UM STABLE FLOW

Minimum stable flow can bedefined 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 an d an increase in NPSHR.

Whats wrong with a drooping

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

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

column or trappedgas.

3. Something mustupset the system tomake it start hunt-ing, i .e . , s tar t ingano t he r pum p i nparallel or throttlinga valve.

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

Minimum f lowbased on the shape ofthe performance curveis not so much a func-tion of the pump as it isa function of the system

w her e t he pum p i splaced. A pump in a sys-t em where the abovecr i t e r i a 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 onh li i i f h d i b i

tistage) with integral bearings. These

bearings 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 stages

FIGURE 2

Recirculation zone s are always on the pres-sure side of the v ane. A show s dischargerecirculation (the front shroud has been leftout for clarity). B shows inlet recirculation.


15/260

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 peripheral

tion, thereby reducing the inlet loss-es. A couple of factors become entan-gled when this i s done. A largerpump means operating back on thepump curve. Minimum flow mu st 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, possibly

10 15 20 25 30 35 40

impeller design. The problem is theresul t 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 anddocum en t . M any s t ud i e s on t hetopic 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 us ed w i t h a m i n i m umamount of information. An exampleof Fraser calculations, which show

the 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 bestefficiency 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 = num ber of impeller vanes

h1 = hub diameter (h1 = 0 for sin-gle suction pum ps)

D1 = impeller eye diameterD2 = impeller outside diameterB1 = impeller inlet widthB2 = impeller outlet widthR1 = impeller inlet radiusR impeller outlet radius

FIGURE 3

Incipient recirculation. Minim um flow is approximately 50% ofincipient flow, w hile minim um interm ittent flow is approximately25% of incipient flow. See text under Recirculation Calculationsfor details

Cm2U2

Discharge Angle 2 Inlet Angle 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

D1

B1

B2

h1

.08


16/260

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

As previously ment ioned,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 r ec i r cula t ion wi th yourpump manufacturer before pur-chas ing a pump, in order toreduce the possibility of problemswith your pump and system afterinstallation and start-up.

SUMMARY

Minimum 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.

REFERENCES

1. F.H. Fraser. Recirculation in cen-trifugal pumps. Presented at the

ASME Winter Annual Meeting(1981).

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

3. J .J. Pau gh. He ad-vs-capacitycharacteristics of centrifugalpum ps . Chemical Engineering(1984).

4. I. Taylor. NPSH still pump appli-cation problem. The Oil and Gas

Journal (1978).

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

Terry W old has been the engi-neering manager for Afton Pumps

for the last four years. He has beeninvolved in pump design for 25

years. M r. W old 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.


17/260

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 conditions

of 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 d ifferences 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 areass igning 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 ofi l l hi h ill

Effects of OversizingBY: IGOR J. KARASSIK

O

Pum p H-Q curve superim posed on system -headcurve

intersection of itshead-capacity curvewi th the system-head curve, as long

as the avai lableNPSH is equal to orexceeds the requiredNPSH (Figure 1).To change this op-erating point in anexisting installationrequires changingei ther the head-capacity curve orthe sys tem-headcurve, or both. Thefirst can be accom-plished by varyingthe speed of thepump (Figure 2), orits impeller dia-meter whi l e thesecond requiresaltering the frictionlosses by throttling

a valve in the pumpdischarge (Figure3). In the majorityof pump installa-tions, the driver isa constant speedmotor, and chang-ing the system-headcurve is used tochange the pump

capacity. Thus, ifwe have providedtoo much excessmargin in the selec-t ion of the pumphead-capacity curve,the pump will have

i h

FIGURE 1

FIGURE 2

Varying pump capacity by varying speed

FIGURE 3


HANDBOOK

HQCurveSystem-

Head Curve

Capacity

Head

Head-CapacityatFullSpeed(N1)

Head-CapacityatFullSpeed(N2)Head-CapacityatFullSpeed(N3) H3

H2

H1

System-Head Curve

FrictionLossesStaticPressur eor Head

}Head

Capacity Q3 Q2 Q1max

Head-CapacityatConstantSpeedH3

H

System-Head Curve


18/260

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

EXAMPLE

Lets use a concrete example

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.,wi th a head-capaci ty 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 bhp

required with our first overly con-servative selection. This is a veryrespectable 12% saving in powerconsumption. Furthermore, weno 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 be

even 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 would

operate at 2820 gpm and take148 bhp. We could be saving morethan 15% in power consumption.Tables 1 and 2 tabulate thesesavings.

And our real margin of safetyis actually even greater than I haveindicated Remember that the fric

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 a

Effect of oversizing a pump

B

A

D

C

H-Q 1800 R.P.M.

143/4"Impeller

Q14

3 /4"Impe

ller

Q

143 /4"

Impe

ller

14"Impeller

Q14

"Impe

ller

Q1

4"Imp

eller

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

fficiency

FeetTotalHead

FIGURE 4


19/260

C l e a r l y ,impor tant energys av i ngs can beachieved if, at thetime of the selec-tion of the condi-t ions of s erv ice ,r e a s o n a b l erestraints are exer-c i sed to avoidi n c o r p o r a t i n gexcess ive safe tym ar g i ns i n t o t herated conditions ofservice.

EXISTINGINSTALLATIONS

But what of existing installations

in which the pumpor pum ps haveexcessive margins?Is i t too late toachieve these sav-ings? Far from it! Asa matter of fact, it ispossible to establishmore accurately thet rue sys tem-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 the

necessary reduced diameter canbe ordered from the pump man-

ufacturer . The or iginalimpeller is then stored for fu-ture use if friction losses areultimately increased w ith 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 impel l er 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 w hich excess ivemargin had originally beenprovided. Again, the originalimpeller is put away for possi-ble future use. s

Igor J. Karassik is Se nior

Consulting Engineer for Ingersoll-

Dresser Pum p Com pany. He hasbeen involved with the pump industry

for m ore than 60 year s. M r.

Karassik is Contributing Editor -

Centrifugal Pumps for Pumps andSystems Magazine.

187.5 bhp. A pump with only a14 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 270 0 GPM

Impeller 143/4" 14"Curve B CBHP 165 145Savings 20 hp or 12.1%

TABLE 1. COM PARISON OF PUM PS WITH 1 43/4 IN. AND14I N. IM PELLERS, W ITH THE SYSTEM THROTTLED

Unthrottled, on Curve A

Impeller 143/4" 14"GPM 3150 2820BHP 177 148Savings 29 hp or 16.4%

TABLE 2. COMPARISON OF PUMPS WITH THESYSTEM 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 INSYSTEM WITH NEW PIPE AND RESULTINGSAVINGS NEW PIPE (UNTHROTTLEDOPERATION, CURVE D)



20/260

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

often necessary to account for the

effect 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 have

viscosities high enough to increasethe f r ict ion drag on a pumpsimpellers significantly. The addi-t ional horsepower required toovercome this drag reduces thepumps efficiency. There are sev-eral analyt ical and empir icalapproaches 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 anaccura te v i scos ity nu mber onwhich to base our estimates. Theviscosity of most liquids is strong-ly influenced by temperature. Thisrelationship is most often shown

by plotting two points on a semi-logarithmic grid and connectingthem with a straight line. The r ela-tionship is of the form:

= AeB/T

where

Fluid Viscosity Effects on Centrifugal PumpsBY: GUNNAR HOLE

WFIGURE 1


HANDBOOK


21/260

NON-NEWTONIANThese are fluids where the

shear rate-shear stress relationshipis nonlinear. They can be dividedinto four categories: Bingha m-plastic f luids are

those in which there is noflow until a threshold shearstress is reached. Beyond thispoint, viscosity decreases with

increasing shear rate. Mostslurries have this property, asdoes Americas favorite veg-etable, catsup.

Dilatant fluids are those ofwhich viscos i ty 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 rate

may 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:

Dynam ic or absolute viscosityis usually measured in termsof centipoise and has the units

of force time/length

2

. Kinem atic viscosity is usually

measured in terms of centis-tokes or ssu (Saybolt SecondsUniversal) . I t is related toabsolute viscosity as follows:

kinematic viscosity

FIGURE 2


22/260

The process of determiningthe effect of a fluids viscosity on

an 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:

mechan ical losses

impeller losses

leakage losses

disk friction losses

Of all external mechanicallosses, disk friction is by far themos t impor tant , according toStepan of f. Th i s i s par t i cu lar -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

The explana-t ion fur ther de-scribes the motionof f lu id in theimmediate n eigh-borhood of thespinning impeller.There Stepanoffmentions the exper-imental results ofothers demonstrat-ing that, by reduc-ing the clearancebetween the s ta-tionary casing andthe impeller, the re-quired power canbe r educed. Healso writes about

the details of some investigationsthat demons t r a te the benef i c i a l

effect of good surface finishes onboth the stationary and rotating sur-faces. Included is a chart preparedby Pf leiderer , based on work byZumbusch and Schul tz-Grunow,that gives friction coefficients forcalculating disk friction losses. Thechart is used in conjunction withthe following equation:

(hp)d = KD

2

u3

where

K = a constant based on the Reynoldsnumber

D = impeller diameter

= 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 f low 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 d iagram from thepaper Engineering and SystemDesign Considerations for PumpSystems and Viscous Service, byC.E. Peter sen , presented a tPacif ic Energy Associat ion,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 n ozzle1)

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 H ydra ulic

Institute Standards. This technique

is based on the use of two nomo-grams on pages 112 and 113 of the14th edition (Figures 71 and 72).They are r eproduced here asFigures 1 and 2. They are intended

WaterCurve-BasedPerform ance % 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 27

Viscous (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 PERFORM ANCE

Note: Pumped fluid specific gravity = 0.9

TABLE 2 POLYNOMIAL COEFFICIENTS


23/260

for use on pumps wi th 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 fol lowing example istaken from pages 114-116 of the

Hydraulic 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 error

of 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 Institute

Standards. 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 + Dx2 P + Dx3 P2 + Dx4 P3 +

Dx5P4 + Dx6P5

where

Cx is the correction factor that must be

applied to the term in question

Dxn are the polynomial coefficients listed

in Table 2

P is the pseudocapacity term definedabove

For comparison, the correctionfactors for the example above (tabu-lated in Table 7 of the Hydraulic

Institute Standards) and those calculat-ed using the polynomial expressionsabove are listed in Table 3.

The problem of selecting a pumpfor use in a viscous service is relative-ly simple once the correction coeffi-

cients have been calculated. If, forexample, we had been looking for apump that could deliver 100 feet of

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

Hwater = Hviscous service/CH1.0

Qwater = Qviscous service/CQ

The next s tep would be tofind a pump having the requiredper formance on water . Af terdetermining the efficiency of thepump on water, we would correctit for the viscous case as shownabove:

viscous service = water x CThe 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 viscous service)

As with water service, thehorsepower requirements at off-

design conditions should alwaysbe checked. s

Gunnar Hole is a principal inTrident Engineering, Inc. in

Houston, 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 a

Registered Professional Engineer inTexas.

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

Per Table 7 of HI Standards 0.635 0.95 0.96 0.94 0.92 0.89

Per Polynom ial Expressions 0.639 0.939 0.958 0.939 0.916 0.887

TABLE 3. CORRECTION FACTOR COM PARISON



24/260

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 common

balancing 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 ar t icle, I includepar t ial ly and ful ly assembledpum p shaft/sleeve/impeller as-semblies as well as individualpump components installed onbalancing machine arbors in thisdefinition.

The three major criteria usedwil l be refer red to as theUnbalanced Force Method (UFM),the Specified Eccentricity Method(SEM), and the Specified CircularVelocity Method (SCVM).

In the UFM the al lowableunbalance permitted in a rotor is

the 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 be

Unbalanced Specified SpecifiedForce Eccentricity Circular

Method Method VelocityAs per API 610 As per Method

6th Edition AGMA 510.02 As per API 610

7th Edition

Residual Unbalance

(RUB), in.oz 56347 Wj 16Wj 4 Wjwhere: N2 NWj = rotor weight per

balance plane, Ibf

N = rpm

= eccentricity, in.Eccentricity () orSpecific Unbalance

in.oz/lbm 56347 16 4N2 N

in.lbm/lbm 3522 0.25N2 Nwhere RUB =Wj see Table 2Unbalance Force (UBF),

lbf where:

UBF =M2 0.10 Wj WjN2 WjNand M = Wj/386 lbfs2/in 35200 140800

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

num ber of balancing standardshave been developed over the

years 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 flexiblecoupl ing vendors . I t 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 peripheral

speeds, the stresses in the rotorand bearings are the same. Thismethod is descr ibed 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 w ith 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 and

length. 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.

ECCENTRICITY

Pump Balancing CriteriaBY GUNNAR HOLE

T

TABLE 1. BALANCING CRITERIA


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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 connecting

them 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 i t back together you take thechance of changing its balance condi-tion. As written, API 610 7th Editionallows a coupling to have a specific

residual 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-mar i ly intended for use on r igidrotorsthose op erating at speedsunder thei r f i rs t cr i t ical 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 U nbalance.

The basic concepts of rigid and

Note: AGMA 515.02 refers to several Balance

Quality Classes. They are sum marized as follow s:

Equivalent I SOAGMA Balance Quality GradeClass , -in. 1,800 rpm 3,600 rpm

8 4,000 19.2 38.39 2,000 9.6 19.2

10 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 th e 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, where

API 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-

Appendix I of API 610 7thEdition briefly discusses some ofthe implications of operating arotor near a critical speed. Theguidelines given there recom-mend separat ion 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 or

resonance 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, , to account for theeffect of sys tem dynamics .Algebraically, the physics of the

situation can be represented asfollows:x = X sin (t )

(/n)2

X = ([1 (/n)2]2 + (2/n)2)0.5

2/n = tan1

1 (/n)2

wherex is the displacement of a point on

the rotor

X is the magnitude of the vibrationat that point

is the mass eccentricity is the operating speed or fre-

quency of the rotor

is the phase angle by which theresponse lags the force

is the damping factor for themode of vibration under consid-eration

X/ is the amplification factor/n is the ratio of operating speed

TABLE 2. BALANCE QUALITY CLASSES



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ntifriction bearings, whichcan utilize either balls orrollers, are u sed 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-r ent ly u t i l i zed in des ign .However, because electric motorsare used more often to drive cen-trifugal pum ps, our discussionwill be based on bearings typical-ly used in quality motors. These

bearings 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 oft ime with minimum frictionand maintenance

m a in t a in c r it ica l t ol er ancesbetween rotating and stationaryassemblies to prevent contactand wear

transmit all variable radial andaxial loads in all operating con-ditions, which include reverserotation startup shutdown

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 show s 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 installer

should 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.

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 of

one 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 should

Bearing BasicsBY RAY RHOE

A

Photo 1. Typical radial bearings


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side, hidden from the installer.Marking the face of each bearingallows the installer to see where the

BAMs 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 be

loaded in either direction. If thisoccurs, the balls in a DACB withno pre load 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, bearingmanu facturers provide bearingsthat have a predetermined clear-ance between either the inner orouter races. Face-to-face bearingshave this clearance between theouter races. When the bearings areclamped together at installation( the outer races are clamped

cause the ro tor loads to changedirection or be eliminated, the bear-ing balls will still be loaded and ball

skid should not occur.One disadvantage of using pre-

loaded bearings is that bearing lifewill be reduced due to the increasedloading. Preloaded bearings shouldnot be used unless design conditionsrequire them.

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

BEARING INSTALLATIONOnce the proper bearings have

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

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

1. RADIAL BEARINGS

To install radial bearings, theyshould be heated in a portable oven

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

2. DACBs

Installation of DACBs followsthe some procedure, except thatadditional care must be taken toposition the bearings properly,line up the burnished alignmentmarks, and not erase the indeliblemarks added on each bear ingface. After the first bearing hasbeen installed, rotate the rotor (ifnecessary) so the alignment markon the inner race is at 12 oclock,then rotate the outer race so it toois at 12 oclock. Before proceedingwith the second bearing, mentallywalk through the procedure .

Remember which direction theface goes and that the burnishedalignment marks must be in thesame position as the first bearingmarks. Also remember you haveabout 10 seconds before the bear-ing seizes the shaft.

The purpose of aligning thefour burnished alignment marksis to minimize off-loading (fight)

and radial runout loads that willoccur if the true centers of thebearings are not lined up. Minorimperfections will always occur,and they must be minimized.Failure to align the marks willresul t in the bear ings loadingeach other.

DACBs come only inmatched pairsthey must be used

together. To verify that a pair ismatched, check the serial numberon the bear ing halvestheyshould be the same, or properlydesignated, such as using bearingA and bearing B.

NEW BEARING RUN-I N

Photo 2. Tw o pairs of DACBs, with the pair on the left positionedface-to-face, th e p air on the right b ack-to-back


28/260

unit at the design rating point andrecord bear ing t empera turesevery 15 min. Bearing tempera-tures should increase sharply andthen slowly decline to their nor-mal operating temperature, usual-ly 2060F above ambient.

D ur i ng the hea t r un ca r eshould be taken to ensure thatthe temperature does not exceedthe value specified by the manu-fac turer . I f i t does , the uni tshould be secured and allowed tocool to within 20F of ambient,or for 2 hours. The unit may thenbe restarted and the test repeatedas necessary until the bearing

temperature peaks and begins todecline.If, after repeated attempts,

bearing temperatures do not showsigns of stabilization, too muchgrease may be present. The bear-ing should be inspected and cor-

i i k

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

3. Bearings must be installed in aclean environment . Contami-

nation is a leading cause of pre-mature failures.

4. Do not pack the bear ing andbear ing cups fu l l of grease .Excessive grease will cause over-heat ing and bal l sk id . SeeExcessive Lubrication underBearing Failure.

BEARING FAILURESFai lure to fol low these four

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

True Brinelling (failure to fol-low Rule 1): This type of bearingf il h i b

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

Contamination (failure tofollow Rule 3): Contamination ofbearings almost always occursduring installation, but can alsooccur when liquids or other con-

stituents from the pump leak orare present in the surroundingenvironment. If contamination isfound in a new bearing beforeinstallation, the bearing should becarefully cleaned and repacked.Evidence of solid contamination inused bearings usually appears asvery small, flat dents in the racesand balls.

Excessive Lubrication (fail-ure to follow Rule 4): Too muchgrease in a bearing may cause theballs to plow their way throughthe grease, resulting in increasedfriction and heat. If the bearingsand bearing caps are packed fullof grease, ball skid could occur.When it does, the balls do not roll,but actually slide against the races.

Experience shows that the bear-ings may be permanently damagedafter more than five minutes ofball skid. Finding packed bearingsand bearing caps is a good indica-tion that too much grease causedthe bearing to fail. Bearing manu-facturers usually recommend thatbearings have 25-50% of their freevolume filled with grease.

Excessive Heat: Failure toprovide adequate heat transferpaths, or operating the componentat excessive loads or speeds mayresult in high operating tempera-tures. Evidence of excessive tem-pera ture usual ly appear s as

Photo 3 . Radial bearing disassembly


29/260

across the country and not cush-ioned from a rough road surface.The load of the rotor is passedthrough the bearing balls, whichwear away or indent the races.Evidence of false brinelling lookssimilar to true brinelling, but maybe accompanied by signs of corro-

sion where the grease film has notbeen maintained. Correction sim-ply involves protecting the unitfrom excessive vibration and usingspecial ly formulated greaseswhere past experience demon-strates the need.

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

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

1. Partially fill a clean bucket orcontainer with clean dieselfuel or kerosene.

2. Insert a clean filter bag into thekerosene conta iner . Thi sensures that no contaminationf rom the conta iner or thekerosene gets into the filter

bag.

3. Using a clean brush, wash thegrease and contamination outof the bearing. The grease will

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

RADIAL BEARING DISASSEMBLY

After removing the grease andany contam ination, you should disas-semble radial bearings by removing

any seals or shields, which are oftenheld in place by snap rings. Then, toremove a meta l r e t a iner , dr i l lthrough the rivets and remove bothretainer halves. Then the bearingshould again be flushed (in a differ-ent location) to remove any metalshavings that may have fal lenbetween the balls and races whendrilling out the rivets. If the bearing

does not freely turn by hand, somemetal par t icles are s t i l l t rappedbetween the balls and races.

Next, place the bearing on thefloor as shown in Photo 3 with theballs packed tightly together on thetop. Insert a rod or bar through the

DACB DISASSEM BLYTo disassemble DACBs, sup-port the face of the outer race andpress down against the inner race.The back of the bearing must beon top.

HANDLING, TRANSPORTATION,AND STORAGE

Common sense appl ies inhandling, storing, or transportingprecision bearings. They shouldnot be dropped or banged. Theyshould be transported by hand incushioned containers, or on theseat of vehiclenot in a bike rack.They should be stored in a cool,clean, dry environment.

Because nothing lasts forever,including bearing grease, bearingsshould not be stored for more than

a few years. After this, the greasedegrades and the bearings maybecome corroded. At best, an oldbearing may have to be cleanedand repacked, using the correcttype and amoun t of grease.

MAINTENANCERoutine maintenance of bear-

ings usually involves periodic

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

To maintain pumps and dri-vers that are secured for longperiods of time, simply turn therotor 1015 revolut ions ever ythree months by hand. This willensure than an adequate greasefilm exists to prevent corrosion ofthe bearing. If this action is nottaken, the bearings may begin tocorrode due to a breakdown inthe grease film. s

Ray W. Rhoe, PE, has a BSCEfrom The Citadel and 15 years expe-

Zero-leakage m agnetic liquid se aldeveloped to retrofit process pum ps



30/260

BY DAVID CUMM INGS

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

increases. Many utilities are empha-

sizing conservation to reduce thenumber of new generating facilitiesthat need to be built. Utilities haveincreased incent ives to conservepower with programs that emphasizedeman d side conservation. These pro-grams often help fund capital equip-ment r eplacements tha t r educeelectrical consumption. Demand sidemanagement programs make replac-ing old pumping equipment morefeasible than ever before.

DETERMINING PUMP EFFICIENCYThe efficiency of a pump p is

ratio of water horsepower (whp) tobrake horsepower (bhp). The highestefficiency of a pump occurs at theflow where the incidence angle of thefluid entering the hydraulic passagesbest matches with the vane angle.

The operat ing condit ion w here apump design has its highest efficien-cy is referred to as the best efficiencypoint (BEP).p = whp/bhp

The water horsepower (whp)may be determined from the equa-tion:whp = QHs/3,960wher e Q = capacity in gallons per

minute, H = developed head in feet,and s = specific gravity of pum pedfluid.

Preferably, the brake horsepow-er supplied by a driver can be deter-mined us ing a t ransmiss iondynameter or with a specially cali-

Twhere ehp = elect r ical power inhorsepower.

EFFICIENCY LOSSESPump efficiency is influenced by

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

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

more uniform hydrau lic passages.Separation of flow occurs w hen a

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

through the passage.Mechanical losses in a pump are

caused by viscous disc friction, bear-ing losses, seal or packing losses, andrecirculation devices. If the clearancebetween the impeller and casing side-wall is too large, disc friction can

and can have s ignif icant powerrequirements.

Internal leakage occurs as theresult of flow between the rotatingand stationary parts of the pump,

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

EXPECTED EFFICIENCIESThe expected hydraulic efficien-

cy of a pump design is a function ofthe pump size and type. Generally,the larger the pump, the higher theefficiency. Pumps that are geometri-cally similar should have similar effi-ciencies. Expected BEPs have beenplotted as a function of specific speed

and pump size. A set of curves thatmay be used to estimate efficiency isprovided in Figure 1. The specificspeed (Ns) of a pump may be deter-mined from the equation:Ns = NQ0.5/H0.75

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

Us ing a pump per formance

curve, the highest efficiency can bedetermined and the specific speedcalculated using the head and capaci-ty at that point. Using the specificspeed and the pump capacity, theexpected efficiency can be estimated.If the pump has bearings or seals that

Centrifugal Pump Eff ic iency

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and motor can be determined. Thecalculated power consumption can becompared with an existing installa-t ion to de termine the value of improving pump performance orreplacing the unit.

EXAM PLE CALCULATION OFPUMP EFFICIENCY

A single-stage end-suctionprocess pump will be used as an

example for an efficiency calculation.The pump uses a mechanical seal andan angular contact ball bearing pairfor thrust. The pumped fluid is waterwith specific gravity of 1.0. Thepump operates at its BEP of 2,250gpm, developing 135 feet of head.The pump speed is 1,750 rpm (note:with the new motor the speed maychange, but to simplify the example it

will be assumed the new and oldmotor both operate at 1,750 rpm).The expecte

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