? HANDBOOK PUMP AND SYSTEM TROUBLESHOOTING

PUMP AND SYSTEM TROUBLESHOOTING

HANDBOOK?

The Pump Handbook Series 1

BY ROSS C. MACKAY AND JOHN T. DAVIS

The article entitled”Pump Suction Condi-tions“ in the May 1993issue of Pumps andSystems stated that,

”The suction of a pump should befitted with an eccentric reducerpositioned with the flat sideuppermost.“ However, somepapers seem to indicate that theremay be an exception to this rulewhen the pump suction is comingfrom an elevated tank. Can youprovide a clarification of thispoint and an explanation of thereasons for the positioning of areducer in the cases of drawingfrom a sump, an elevated tank,and a side suction condition?

David W. Lawhon, RotatingEquipment Engineer & Jeffrey A.Robinette, Project Engineer BasicChemicals Group Occidental ChemicalCorporation, Ingleside, TX

The physical location of apump relative to its sup-ply source does notchange the rules of inletpiping design, and the rea-

son for their existence. When design-ing the layout of suction piping, it‘snot necessary to be influenced bywhether or not it‘s approaching thepump from the side, the top or thebottom. The only concern over pipingconfiguration is the need to deliverthe liquid to the eye of the impeller ina smooth parallel flow of uniformvelocity.

An inverted eccentric reducerand a concentric reducer produce thesame problem in a horizontal line(Figure A).

The flow through an elbow in ahorizontal and a vertical line, createsa similar uneven flow pattern (FigureB).

Some consideration has beengiven to the possibility that an accept-able flow pattern could be achievedby leading a vertical line, downthrough the elbow, into a horizontaleccentric reducer in the inverted posi-tion. However, when we consider the

combination of these, it is evi-dent that it doesn’t really help(Figure C).

If it is absolutely essentialthat an elbow and a reducerbe adjacent to each other,close to the pump suction,consider combining the twointo a reducing, wide sweepelbow (Figure D).

Severe space restriction isthe only sound reason for notproviding suitable inlet pipingfor centrifugal pumps and, ifyou’re in this position, youmay be faced with somechoices. At that point, youmay wish to reconsider thefive rules previously outlined:1. provide sufficient NPSH2. minimize friction loss3. no elbows on the suction

flange4. eliminate vapor from the

suction line5. ensure correct piping align-

ment

Any compromise of theserules will reduce pump reliabil-ity, while conformance to themwill ensure that inlet pipingdesign will not be a factor inany future pump problems.

When selectingan electric motorfor a centrifugalpump, is there anadvantage to pur-

chasing a motor with anefficiency of 96% in lieu ofa motor with an efficiencyof 87% if the pump operat-ing efficiency is only 75%?

Albert M. Diaz, ProductionEngineer Advisor, Unocal, EnergyResources Division, Midland, TX

The efficiency of the elec-tric motor driving a pumpor any other load willdirectly affect the efficien-

cy of the pump/ motor system. Thefinancial advantage gained by select-ing a motor higher in efficiencydepends on numerous factors,including the cost of electricity, thehours of operation, the load imposedon the motor, and the differential ini-tial cost of the motors being com-pared. In a motor retrofit, analysis of

Inlet Piping Configuration and Motor Efficiency

Q:

A:

Q:

A:

FIGURE 1

Fig. A

Fig. B

Fig. C

Fig. D

Air Pocket

Air Pocket

Air Pocket

2 The Pump Handbook Series

the operating speed of the systemshould be considered to avoid havingthe higher efficient motor consumemore power, producing additionaland possibly unnecessary processoutput.

Simplistically speaking, motorefficiency multiplied by pump effi-ciency determines the system effi-ciency. There are other losses toconsider, but given the motor andpump efficiencies in your question,the system efficiencies would be72% versus 65.3%. This system effi-ciency differential might require con-siderable additional expenditure. A

common rating used on a horizontalinjection pump is 300 hp Two Pole.

Given• 7200 hours per year operation

• 6 cents per KW-hr power cost

• 300 hp motor

• operated at full load continu -ous

Calculated energy costs per year:motor (96.2% efficient) $100,500

motor (91.0% efficient)$106,243

The savings is $5743 per year!This is probably a worst case analy-sis: running all year, expensive ener-gy, large efficiency differential—butobviously a significant amount ofenergy costs are involved and it iscommon for the extra cost of a higherefficient motor to be justified by ener-gy savings. Analysis that includescash flow, tax rates, power factors,and other issues can be performed fora more complete review. ■

Answer to first question provided byRoss C. Mackay, second answer byJohn T. Davis.


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

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

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

CAVITATION EFFECTS

BUBBLE FORMATION PHASEFlow is reduced as the liquid

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

BUBBLE COLLAPSE PHASE1. Mechanical damage occurs as

the imploding bubbles removesegments of impeller material.

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

NET POSITIVE SUCTION HEADWhen designing a pumping

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

Cavitation and NPSH in Centrifugal PumpsBY PAUL T. LAHR

C

FIGURE 1

Vapor pressures of various liquids related to temperature.

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

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

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

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

NPSHR is always referencedto the pump impeller center line.


HANDBOOK??

1000

800

600500

400

300

200

100

80

605040

30

20

10

8

654

3

2

1.0

.80

.60

.50

.40

.30

.20

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

28"

28.5"

29"29.1"29.2"29.3"

29.4"

29.5"

29.6"

29.7"29.72"

10"

15"

20"

22.5"

25"

26"

27"

80

60

50

40

30

20

14

1052 05

985

800

600500

400

300

200

140

100

-60° to 240°F

TEMPERATURE–F

CARBON DIOXIDE

NITROUS OXIDE

ETHANE

MONOCHLOROTRIFLUOROMETHANE

HYDROGEN SULFIDE

PROPYLENE

PROPANE

AMMONIA

CHLORINE

METHYL CHLORID

ESULFUR D

IOXID

E

ISOBUTANE BUTANE

ETHYL CHLO

RIDE

METHYL FORMATE

DIETHYL E

THER

METHYLENE C

HLORID

E

DICHLO

ROETHYLENE

ACETONE

DICHLO

ROETH

YLENE (C

IS)

CHLOROFORM (TRIC

HLOROMETHANE)

CARBON TETRACHLO

RIDE TRIC

HLOROETHULENE

WATE

R

HEAVYWATER (S

P.GR. A

T 70 F

=1.106

GA

UG

E P

RE

SS

UR

E–L

BS

. PE

R S

Q. I

N.

VA

CU

UM

–IN

CH

ES

OF

ME

RC

UR

Y

AB

SO

LUT

E P

RE

SS

UR

E–L

BS

. PE

R S

Q. I

N.


It is a measure of the pressuredrop as the liquid travels from thepump 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, droppingat generally 20–30% of the bestefficiency flow, then increasingwith flow. The complex subject ofturbulence and NPSHR at lowflow is best left to another discus-sion.

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 the

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.

FIGURE 24a SUCTION SUPPLY OPEN TO ATMOSPHERE-with Suction Lift

CL

Ls

PB NPSHA=PB – (VP + Ls + ht)

4b SUCTION SUPPLY OPEN TO ATMOSPHERE-with Suction Head

NPSHA=PB + LH – (VP + ht)

PB

LH

CL

4c CLOSED SUCTION SUPPLY -with Suction Lift

NPSHA=p – (Ls + VP + ht)

p

CL

Ls

4d CLOSED SUCTION SUPPLY -with Suction Head

NPSHA=p + LH – (VP + ht)

p

CL

LH

A B C D E

E

A B C

D

ENTRANCELOSS

FRICTION

TURBULANCEFRICTION

ENTRANCELOSS AT

VANE TIPS

INCREASEPRESSURE

DUE TO IMPELLER

PO

INT

OF

LOW

ES

T P

RE

SS

UR

EW

HE

RE

VA

PO

RIZ

AT

ION

ST

AR

TS

INC

RE

AS

E P

RE

SS

UR

E

FIGURE 3

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

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

SUCTION SPECIFIC SPEEDThe concept of suction specif-

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

N x Q0.5Ss = ——————

(NPSHR)0.75

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

best efficiency pointNPSHR = NPSHR at Q with

the maximum impeller diameter

The system designershould also calculate thesystem suction specificspeed by substitutingdesign flow rate and thesystem designer’s NPSHA.The pump speed N is gen-erally determined by thehead or pressure requiredin the system. For a low-maintenance pump sys-tem, designers and mostuser specifications require,or prefer, Ss values below10,000 to 12,000.However, as indicatedabove, the pump Ss is dic-tated to a great extent bythe system conditions,design flow, head, and theNPSHA.

Figures 5 and 6 areplots of Ss versus flow ingpm for various NPSHA or NPSHRat 3,500 and 1,750 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 to1,000 gpm if the maximum Ss is tobe maintained at 12,000. Variousoptions are available, such asreducing the head to allow 1,750

Q1Q2

100% CAP Q3

Q4

NPSH

3%

NP

SH

R

TO

TA

L H

EA

D

FIGURE 4

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

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

198

7

6

5

4

3

2

198

7

6

5

4

3

2

1

HSV=2

3

4

5

678

910

HSV=12

14 16 18 20

28

3236

40

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

FIGURE 5


S 1Su

ctio

n sp

ecifi

c sp

eed


rpm (Figure 7). Thiswould allow flows to4,000 gpm with 20 feetof NPSHA.

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 bewell established andunderstood. A is the nor-mal operating point. B isthe maximum flow for

cavitation-free operation. C is theminimum stable flow, which is dic-tated by the suction 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 willnot create problems. However, ifthe pump is to be operated atreduced flow, then the Ss valuemust be given careful considera-tion. ■

REFERENCES1. Goulds Pump Manual.

2. Durco Pump Engineering Manual.

3. Hydraulic Institute TestStandards—1988 CentrifugalPumps 1.6.

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

HE

AD

NP

SH

- F

EE

T

GPMC A B

4

3

2

1

FIGURE 7

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

198

7

6

5

4

3

2

198

7

6

5

4

3

2

1

HSV=12

HSV=1

1098

6

3

2

14

45

7 1618 20

2832 36 40

50

HSV=24

HSV=45

Solution for

S=N

for N=1,750 rpm

Q Hsv

0.75

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

FIGURE 6

S 1Su

ctio

n sp

ecifi

c sp

eed


HANDBOOK??


BY STEPHEN D. ABLE

INTRODUCTIONIn cases where suction perfor-

mance is evaluated in terms of suc-tion lift, actual pump performancemay not meet the user’s expectation.Suction lift is generally sufficientwhen handling nonvolatile liquids(with a specific gravity near 1.0) fromopen supply tanks at ambient temper-atures. If the supply is pressurized,the fluid is volatile, or has a specificgravity other than 1.0; the use of netpositive suction head (NPSH) wouldbe indicated. Several examples arepresented which illustrate the need toevaluate pump suction performancein terms of NPSH.

Confusion can arise concerningthe meaning and use of “priming”verses suction lift, and the meaningof “dry” and “wet” suction lift.

NPSHAAccording to Hydraulic Institute

Standards (Ref. 1), net positive suctionhead available (NPSHA) is the totalsuction pressure, including allowancefor acceleration head available fromthe system at the pump suction con-nection, minus the vapor pressure ofthe liquid at the pumping tempera-ture. NPSHA for a reciprocating pumpis normally expressed in pounds persquare inch (psi) or feet. If the pres-sures and temperatures are providedat some distance from the suction ofthe pump, frictional head losses mustbe calculated and subtracted.

NPSHA = (P – Pv – Pf) x 2.31 sp. gr.-H

whereP = absolute pressure on the surface of

the liquid (psia, i.e., atmosphericpressure)

Pv = vapor pressure of the liquid (psi)Pf = pressure loss due to inlet piping fric-

tionsp gr = specific gravity of the liquidHs = height of the pumps suction inlet

above the tank level

NPSHRNet positive suction head

required is the total head that the

pump requires to operate properlywithout cavitation (local boiling of thefluid), a reduction in pressure-flowperformance, excessive vibration ornoise. NPSHA must always begreater than NPSHR.

SUCTION LIFTHydraulic Institute defines “total

suction lift” as the pressure belowatmospheric at the inlet port of thepump plus the suction system fric-tional losses. Thus, total suction lift isequivalent to the height of the pumpsuction above the tank level plus thepiping system frictional losses.

Many pump companies use aclosed loop test arrangement withwater to establish the NPSHR of theirpumps as well as the head, flow andefficiency.

PRIMING AND DRY SUCTION LIFTSuction lift (also NPSH) and

priming are two separate issues.Priming has to do with how well thepump can draw a vacuum and cre-ate a suction in the inlet. In the caseof positive displacement pumps, the“dry lift” capability of a pump indi-cates how well the check valveswork when sealing against a vacu-um. If a pump has perfect valvesealing, it could lift a column ofwater 33.96 feet at sea level; that is,14.7 psia x 2.31/1.0 x sp gr.

SPECIFIC GRAVITY AND VAPOR PRESSURE EFFECTS ON SUC-TION LIFT

If the liquid has a specific gravi-ty of 2, the pump could only lift theliquid a maximum of 16.98 feet (that

Pump Suction Lift, NPSH and Priming

FIGURE 1

Suction piping schematic for Example 4, pumping a viscous fluid

Atmospheric Pressure (14.7 psia)+ 5 psig + 19.7 psia

4 inch Pipe(10 feet Total)

GateValve

90° Elbow

5 feet

Entrance


is, 33.96/2). If the liquid has a signifi-cant vapor pressure, it would havethe effect of lowering the differentialpressure across the liquid column. In other words, at sea level therecould no longer be a total of 14.7 psi available to lift the liquidcolumn. If, for example, the liquidhas a vapor pressure of 1 psi and aspecific gravity of 2, a pump withideal valving could lift the liquidonly 15.82 feet; that is, 14.7 psia – 1psia x 2.31/2 sp gr. Reduced atmos-pheric pressure has a similar effect.

PIPING LOSSESIf the pressures and tempera-

tures are provided at some distancefrom the suction of the pump, fric-tional head losses must be calculat-ed and subtracted. To calculate orlook up piping loss informationthrough piping of various sizes andviscosities see Cameron HydraulicData or Thermodynamic Properties ofSteam (Ref. 2 and 3). These refer-ences also provide information onlosses through fittings, elbows, etc.

EXAMPLE 1: LIFTING COLD WATERAn application requires a pump

to lift 20 feet at the flow rate speci-fied. The factory test data indicatethat the pump has an NPSHR of 12feet at that flow rate. Determine theNPSHA of the application and staticheight that the pump could lift thefluid for this application.

Note: The vapor pressure ofwater at 70°F is equal to 0.3631 psi(Ref. 3). Assume that frictional lossesare negligible.

NPSHA = [(14.7 – 0.36 – 0) x 2.31] / (1.0 - 20) = 13.1 feet

Static height from the NPSH equation:

Hs = [(14.7 – 0.36 – 0) x 2.31]/1.0 – 12(NPSHR) = 21.1 feet

Therefore, the pump selectedcan be used in this applicationbecause the NPSHA is greater thanthe NPSHR. The resulting calculationfor the static height indicates this aswell.

EXAMPLE 2: LIFTING A HIGH SPECIFIC GRAVITY FLUID

If the fluid pumped in Example 1 is changed to sulfuric acid (sp gr =1.83), the suction performance wouldchange for this application. Deter-mine the static height, Hs, that thepump would lift.

Note: At 330°C, sulfuric acid hasa vapor pressure of 1 atmosphere(Ref. 4). Because water exhibits thispressure at 100°C, assume that thevapor pressure for sulfuric acid atroom temperature is negligible.

Hs = [(14.7 – 0 – 0) x 2.31]/1.83 – 12(NPSHR) = 6.6 feet

FIGURE 2

Friction factors for commercial pipe (Cameron Hydraulic Data)


Therefore, a pump that can lift21.1 feet on a cold water test can onlylift 6.6 feet of cold sulfuric acid.

EXAMPLE 3: LIFTING COLD SULFURIC ACID

This example is for the samepump as in Examples 1 and 2, but

with 1 psig of frictional line pressureloss (Figure 1).

Hs = [(14.7 – 0 – 1) x 2.31]/1.83 – 12(NPSHR) = 5.3 feet

The effect of the line losses fur-ther reduces the static height of fluid

that can be lifted by the pump. Notealso that atmospheric pressure dropsabout 1 psi for 1000 feet of elevation.Therefore, elevation can affect pumpsuction performance as well as linelosses.

EXAMPLE 4: PUMPING A VISCOUS FLUID

For viscous materials that are ator near room temperature and have asupply tank at least a few feet away,use an NPSH calculation rather thanonly suction lift. Suction lift does nottake into account fluid viscosity andthe resulting frictional losses in thesuction system. For this example:• tank is pressurized to 5 psig

• atmospheric pressure equals 14.7psi

• pumped fluid is at room temper-ature with a viscosity of 100,000cP

• 10 feet of 4-inch pipe on suction

• one 4-inch elbow

• one 4-inch open gate valve

• specific gravity equals 1.0

• pump suction is 5 feet belowminimum tank level

• maximum flow rate duringstroke is 2 gpm

Since it is generally difficult tofind tabular data for friction loss inpiping, valves, and fittings at high-er viscosities the pipe friction losscan be calculated as follows:

Hf = f x L x V2/D/2/g

Note: This equation is for nontur-bulent (laminar) flow. If Re > 2000flow is considered turbulent youmust use a “Moody Diagram” (Figure2).

where

f = friction factor = 64/Re

Re = Reynold’s number = 3162 x Q/d/k[Ref. 4]

Q = flow, gpm during the stroke of theplunger

d = diameter of pipe, inches

D = diameter of pipe, feet

L = pipe length, feet

FIGURE 3

Resistance of valves and fittings to flow of fluids in equivalentlength of pipe (Hydraulic Institute Standards)


k = kinematic viscosity, centistokes = v,

absolute viscosity (centipoise) x spe-

cific gravity

V = velocity of fluid, during the stroke of

the plunger = 0.408 x Q /d2 [Ref. 4]

g = gravitational constant, 32.2 ft/s2

So:

Re = 3162 x 2/4/100000 = 0.0158

V = 0.4085 x 2/42 = 0.05 ft/s

f = 64/Re = 64/0.0158 = 4050

Hf = 4050 x 10 x (0.05)2/(4/12)/2/32.2 =4.7 ft

= 4.7/2.31 = 2.04 psiAvoid using the head loss formu-

la (hf = k x V2 / 2g) because it is notcorrected for viscosity effects. Thefriction losses at the tank exit, valves,and fittings are given by looking upthe equivalent length of straight pipefor the restrictions (Figure 3):

1. L (4-inch elbow) = 10.1 ft of pipe

2. L (open gate valve) = 2.3 ft ofpipe

3. L (tank exit is the ordinary exit)= 6 ft of pipe

The equivalent length of straightpipe for the three restrictions is equalto 18.4 ft of 4-inch pipe (10.1 + 2.3 +6).

The head loss through the threerestrictions can be found by knowing

the loss in 10 feet of straight pipe.From the earlier calculation, there isa 4.7 foot head loss in 10 feet ofstraight pipe. Therefore, the loss inthe three restrictions is:

hf multiplied by the three restrictions= 4.7 ft x 18.4 ft of pipe

10 ft of pipe = 8.7 ft (or 3.7 psi)

The total frictional losses are:

hf = 4.7 + 8.7 = 13.4 ft (or 5.8 psi)

The NPSHA is equal to:Absolute tank pressure: 5 + 14.7= 19.7 psia (or 45.5 ft)

vapor pressure of the fluid at thepump suction: assume it is zero

piping system losses: 5.8 psi (13.4ft)

height of fluid above pump suc-tion : 5 ft

NPSH A = [ (14.7 + 5) – 0 – 5.8] x(2.31 / 1.0) + 5 = 37.1 ft

In this example the NPSHA isgreater than in the other examplesdue to the addition of tank pressureand the maintenance of a minimumtank level above the pump. Fluidswith a high viscosity often undergochanges in viscosity when set inmotion, which can significantlyaffect the prediction. (Normally, the

pump will perform better than pre-dicted, since most high viscosity flu-ids are shear-thinning).

Depending on pump type, pumpoperation and efficiency can bealtered when handling viscous fluids.■

REFERENCES1. Hydraulic Institute, 1983,

Hydraulic Institute Standards forCentrifugal, Rotary, andReciprocating Pumps, FourteenthEdition.

2. Heald, C.C. (1988), CameronHydraulic Data, SeventeenthEdition.

3. Keenan, J.H. and Keyes, F.G.,(1964), Thermodynamic Proper-tiesof Steam, Thirty Sixth Printing,page 28.

4. The Crane Company (1988),“Flow of Fluids through Valves,Fittings and Pipe,” TechnicalPaper No. 410, page 32.

5. Weast, R.C. (1969), Handbook ofChemistry and Physics, 50thEdition, page D144.

Stephen D. Able is a SeniorEngineering Consultant with AROFluid Products Division, Ingersoll-Randin Bryan, OH.


HANDBOOK??


BY JOHN SIDELKO

Knowing how impor-tant NPSH is to the sat-isfactory operation ofour pumps, my col-leagues always recom-

mend that we provide more thanthe amount cited by the pumpvendor. How much more NPSH isenough?

The Net Positive SuctionHead (NPSH) required,quoted by the pump ven-dor, is based on HydraulicInstitute Standards. It is a

measure of the total energy in thefluid on the suction side of the pump.This value of low suction energy atany given flow and speed causes a3% reduction in the normal differen-tial head created by the pump. Theoverall performance is adverselyaffected because the pump is movingliquid away faster than it is allowedto enter.

This problem occurs more fre-quently when handling liquids closeto their boiling point. Vibration (gen-erally not associated with anymechanical frequency in themachine), shortened bearing and seallife, and possibly even impeller andcase damage will happen repeatedly.Other criteria subject to evaluationand justification during the purchas-ing process such as total installedcost, efficiency, and maintenance, areovershadowed by the absolute neces-sity for enough NPSH.

In low energy pumping systems(low flow, low speed, and low horse-power), even though something iswrong within a pump operating at itsvalue of NPSH required, the amountof energy involved is too low to causedamage. Hydrocarbons, due to theirchemistry, also don’t dissipate muchenergy during cavitation. It is com-mon in these situations to supply twofeet more than the maximum NPSHrequired by the pump. Critical ser-vices that are the heart of the process,and/or high energy machines (highflow, high speed, or high horsepow-er) are the ones that should compelpump users to add a margin of safety

Q: so that system NPSH availablealways exceeds the pump NPSHrequired. Figure 1 will help deter-mine the amount of cushion neces-sary to achieve reliable pumpoperation.

Many pump manufacturers gen-erate diagrams like Figure 1, based onfactory tests, to accurately determinethe Hydraulic Institute definition ofNPSH required over the completerange of speed and flow for eachmodel they sell. Figure 1 clearlydepicts the rate of pump performancedegradation due to decreasing valuesof NPSH. It is this rate of degradationand the amount of energy being addedto the fluid that determines if a pumpwill perform satisfactorily under mar-ginal amounts of NPSH available.

Figure 1 represents additionalinformation that users should requestof pump vendors because it is notcommonly published data. The illus-tration should represent the pump’sbehavior at the most difficult operat-ing point—the maximum rated flowand speed for a specificinstallation.

• Insufficient NPSHA(below the HydraulicInstitute definition ofNPSH required) willcause unreliablepump operation.

• Marginal NPSHA isthe transition zonewhere pumps may ormay not work well.Low energy pumpswill be fine. Criticalservices or high ener-gy machines shouldbe operated withenough systemNPSH available tomaintain 99% differ-ential head or more. Dependingon the inlet geometry of thepump, sometimes only severalfeet of increased NPSH may beneeded to push the pump intothe area of reliable operation.

• Sufficient NPSHA is a safe oper-ating region for any centrifugalpump.

Two situations exist: either newequipment is under evaluation tocomplete a specific process require-ment, or the pumps and system arealready in place.

NEW EQUIPMENTMany articles have been written

and commercial software has beendeveloped to help calculate systemNPSHA. When NPSH available isonly zero to ten feet greater thanNPSH required, the pump vendorshould be asked to provide a graphsimilar to Figure 1 for the pump oper-ating at its maximum rated speed andflow.

Try to select a pump that inter-acts with NPSH available in the sys-tem to ensure 99% differential heador better. If not, contact the pumpvendor to see if any internal hard-ware changes can be made toimprove the suction performance.

Often larger inlet eyediameters are used tolower NPSH required.

Piping alterations onthe suction side of themachine also affectNPSH available. Thepump user must deter-mine this value andmake the necessaryimprovements. Largersuction lines, shorter dis-tance, fewer bends andrestrictions, or higherminimum tank levelhelp raise NPSH avail-able in the system.

EXISTING EQUIPMENTAlthough it is not as

easy to solicit coopera-tion after the equipment sale, manu-facturing tolerances are accurateenough by today’s standards that atypical representation of Figure 1 canbe provided by the pump vendor.Instrumentation should be employedwhenever possible in the field to

How Much NPSH is Enough?

A:

High energymachinesshould beoperated

with enoughsystem NPSHavailable to

maintain 99%differential

head or more.


measure suction pressure, tempera-ture, flow, fluid characteristics, andbarometric pressure. This informa-tion can be applied to the followingequation:

NPSHA = PB – Vp ± GR + HVwhere:PB = barometric pressure converted tofeetVp = fluid vapor pressure converted tofeet

GR = gauge reading at pump suctionconverted to feet

HV = velocity head = V2/2g

V = fluid velocity at pump suction flangein feet/sec

g = gravitational constant

(32.174 ft/sec2)

Plotting the NPSH measured inthe system on Figure 1 will quicklyassess the health of your installation.

As noted above, some hardwareadjustment within the machine mayhelp. Sometimes radical changesmust be made, such as adding aninducer to the pump or changingmanufacturers. System changes tolower the pump, raise the tank,reduce process temperature,increase line size, shorten the suc-tion run, or minimize elbows andrestrictions may easily prove to bemore expensive.

Misunderstanding NPSH can bea costly error. There simply must beenough NPSH available for any

pump installation to operate reliably.■

John E. Sidelko is ProductDevelopment Manager for A.R.Wilfley and Sons, Inc. He serves onthe Pumps and Systems EditorialAdvisory Board.

FIGURE 1

Increasing NPSH available versus percent of pump differential head.


HANDBOOK??


BY DAVID WILLIAM SPITZER

variable speed drive is adevice that changes the speedof a driven load, typically byusing electrical, mechanical,or hydraulic means (Figure 1).

Considering the wide range ofoverlap in their application, it shouldbe no surprise that the use of variablespeed drives often falls between thecracks of the traditional technical dis-ciplines.

A VARIABLE SPEED DRIVE:1. may be an electrical device speci-

fied and installed by an electricalengineer, such as an electricalswitch gear

2. may be a mechanical devicespecified and installed by amechanical engineer, such as agear reducer or hydraulic cou-pling

3. may be a final control element,such as a replacement for a con-trol valve that can affect the con-trol of the process, an areaimportant to the instrumentationand control engineer

4. may affect the electrical powerdistribution system by reducingelectrical demand

5. may affect the operation andproper functioning of themechanical equipment by oper-ating slower, and potentiallyreducing maintenance

6. may affect the operation of theprocess and be the key compo-nent required to improve theprocess because desired opera-tion cannot be achieved with acontrol valve

Knowledgeable generalists whoappropriately rely on specialists (suchas electrical, mechanical, process,instrumentation, and process controlengineers) are in the best position toapply variable speed drive technology- and they are most likely to come upwith viable, technically correct solu-tions. With the increased trendtoward specialization of job functions,however, people with broad knowl-

A edge and appropriate skills are be-coming more difficult to find.

The primary advantage of vari-able speed drive technology is theeconomy gained by reducing electri-cal energy con-sumption; how-ever, other ad-vantages that are difficult toquantify may be achieved inmost applica-tions. Theseinclude reduc-tions in mainte-nance require-ments as a resultof fewer rota-tions, reducedforces on inter-nal componentsand seals, andimproved over-all performanceas a result oftighter control ofthe process. Forexample, a vari-able speed driveinstallation canimprove control

of flow as compared to a controlvalve installation, in that it virtuallyeliminates valve-related hysteresis.Such an installation may significantly

Variable Speed Drives: Advantages and Pitfalls

FIGURE 1 A & B

FIGURE 2

A. VSD application with variable motor speed

B. VSD application with constant motor speed

VariableSpeedDrive

Motor andEquipment

Speed Varies

Equipment SpeedVaries

ConstantMotorSpeed

Full speed operating point

Increasing shafthorsepower

System

Full speed pump curvePump curve at

reduced speed

Reduced speedoperating point

Flow

Pressure

Motor MechanicalEquipment

MotorVariableSpeedDrive

MechanicalEquipment


improve product quali-ty and process efficien-cy.

CONSTANT TORQUE MECHANICAL LOADS

Constant torqueloads, such as thosefound in positive dis-placement equipment,can be analyzed in a rel-atively straightforwardmanner because they require nearlyconstant torque input as speed is var-ied. As such, shaft horsepowerrequirements, equipment capacity,and electrical energy consumptionvary directly with speed.

For example, if a piece of equip-ment must produce 75% of its capaci-ty (at rated speed), it will operate at75% of rated speed, require approxi-mately full torque, and consumeapproximately 75% of its electricalrequirements at rated speed.

CENTRIFUGAL MECHANICAL LOADS

Centrifugal pumps and fans fol-low the affinity laws, which are non-linear, adding complexity to theeconomic analysis of any given appli-cation. Graphical representations ofvariable speed drive operation clearlyshow a reduction in shaft horsepower(which will reduce electrical energyconsumption) while providing therequired hydraulic energy output(Figure 2).

Economic analysis is beyond thescope of this article (see reference).However, it is important to recognizethat the brake horsepower of cen-trifugal loads varies with the cube ofthe speed. So, reducing speed byonly 10% (to 90% of rated speed) willreduce the brake horsepower by27%:

[100(1 – 0.93) ]

TYPICAL APPLICATIONSVariable speed drives typically

alter equipment speed to provide therotational energy input necessary tosupply the hydraulic energy output tothe process. In most situations equip-ment speed is varied to control level,pressure, or flow (Figure 3).

With the exception of headerpressure control, variable speed dri-ves are generally not applicable when

the output from one piece of equip-ment feeds multiple users becausemore than one flow cannot be inde-pendently controlled by only onefinal control element.

In addition, care should be takenwhen attempting to manipulate theoperating speeds of equipment work-ing in parallel to ensure that eachpiece of equipment supplies anappropriate part of the total load.

Whereas many variable speeddrive applications may replace a con-trol valve or damper, the best applica-tions improve the process as well. Inone such application, piping leaks thatwere caused by full-speed operationbecause the pump produced toomuch head were minimized by open-

ing the bypass valve to the feed tank,causing the pump to ride out on itscurve, effectively reducing dischargepressure (Figure 4A). While thisapproach ”worked,“ it also increasedoperating costs significantly becausethe increased recirculating flow (tothe tank) produced no benefit to theprocess. A variable speed drive wasinstalled on the centrifugal pump tocontrol the header pressure, allowingthe bypass valve to be closed (Figure4B). This effectively eliminated thebypass flow and reduced energy con-sumption by 73%. In addition, tightcontrol of the discharge pressure wasachieved under varying operatingconditions that resulted in more uni-form flow to process equipment. It

FIC

Flow

FIC

Flow

FIGURE 3

A. Typical full-speed pump flow control B. Typical variable-speed pump flow control

Supply to Plant

Supply to Plant

Throttled to reducesupply pressure

A. Full-speed operation

Return

Return

Closed

PIC

P

P

Q

Q

B. Variable-speed operation

FIGURE 4 A & B


should be noted that variable speeddrive applications that replace controlvalves typically reduce electrical ener-gy consumption by 30–70%.However, these electrical energy sav-ings may be small compared toprocess-related savings in applicationswhere the variable speed driveimproves the process.

POTENTIAL PITFALLSThe potential pitfalls of applying

variable speed drives to rotating equip-ment include the following:1. Constant torque variable fre-

quency drives must be able toprovide full motor current at allspeeds. To avoid premature fail-ure, the electrical equipmentoperating these loads should bespecified, designed, and rated forconstant-torque operation.

2. In larger applications, electricalequipment should be suitablydesigned to be integrated into theelectrical system to avoid prob-lems such as electrical noise,ringing, and harmonic distortion.

3. The motor must be capable ofdissipating the heat it generatesat all speeds, including lowspeeds, when the cooling fan alsooperates slowly. Solutions to thisproblem include appropriatelylimiting the motor operatingspeed range, upgrading themotor winding insulation, over-sizing the motor to provide moreheat dissipation, utilizing a totallyenclosed non-ventilated (TENV)motor, or installing auxiliarymotor cooling to dissipate heat.This problem is especially preva-lent in constant torque applica-tions where full-load currentflows, even though the effective-ness of the motor cooling fan isreduced or nonexistent.

4. In Division 1 hazardous loca-tions, the National Electric Code

(NEC) requires that the motor beapproved for the hazard. Whilefull-speed motors are approvedfor full-speed operation at maxi-mum load (and maximum fancooling), variable speed drivemotors must be approved overthe entire operating speed range.In centrifugal applications, theworst case is usually at fullspeed, a condition for which themotor was designed. However,in constant torque applications,maximum heat dissipation isrequired at all speeds. Approvedmotors for each type of variablespeed drive load are becomingavailable to meet these require-ments, utilizing appropriate heatdissipation techniques describedin item 3 (as required).

5. The mechanical equipment mustbe capable of operating at reducedspeeds. For example, a liquid-ringvacuum pump is a questionablecandidate for variable speed driveoperation because it fails to pro-duce a vacuum below approxi-mately 80% of its rated speed.

6. The mechanical equipment mustoperate without damage atreduced speeds. Some equip-ment can be damaged by ineffi-ciencies and slip at slower speedsthat cause overheating. Equip-ment can also be damaged whenits lubrication system becomesinadequate at reduced speeds.

7. In most installations variable speeddrives cannot securely stop theprocess fluid—that is, provide tightshutoff. When tight shutoff isrequired for process reasons, acontrol valve can be installed inaddition to the variable speeddrive. In certain applications, acheck valve may be required toprevent reverse flow conditionsthat can compromise safety orupset the process.

8. Consideration must be given tothe fact that a mechanical vari-able speed drive is an additionalpiece of equipment with movingparts subject to the wear andtear, heat and maintenanceneeds of industrial machinery.

9. Hydraulic variable speed drivesinvolve two energy transforma-tions (rotating to hydraulic andhydraulic to rotating) that mayadd to electrical energy costs. Inaddition, leaks and other mainte-nance headaches will sometimesoccur, particularly in high-pres-sure situations.

SUMMARYThis discussion, by no means

complete (Ref. 1), should give readers abasic understanding of some of theadvantages and pitfalls associated witha technology that can be the key to sig-nificant process improvements andenergy savings. While there are a num-ber of pitfalls, some of which are out-lined, the reader should not beintimidated because the great majorityof problems can be solved by utilizingthe knowledge and skill of specialists.An individual familiar with themechanical nature of the equipment isin a unique position to initiate and sup-port the application of variable speeddrive technology when equipmentspeed can be safely and effectivelyreduced. ■

REFERENCED.W. Spitzer. Variable Speed

Drives—Principles and Applicationsfor Energy Cost Savings, 2nd EditionRevised, Instrument Society ofAmerica, 1990.

David William Spitzer, P.E., isManager of Utility and InstrumentationEngineering for Nepera, Inc. inHarriman, NY.



HANDBOOK??BY JAMES NETZEL

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

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

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

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

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

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

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

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

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

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

The maximum head or discharge

pressure of a centrifugal pump can be

easily estimated if the impeller

diameter, number of

impellers used, and rpm

of the driver are known.

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

Q:A:

17

16

14

13

12

11

10

9

8

7

6

5

4

3

2

1

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

FIGURE 1

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

RPM x 1000

Hea

d in

Fee

t x 1

000


What different types ofseal lubrication exist?

A mechanical seal isdesigned tooperate in manytypes of fluids.The product

sealed becomes the lubri-cant for the seal faces.Many times the fluidbeing sealed is a poorlubricant or contains abra-sives that must be takeninto account in the sealdesign. The design of theseal faces, materials ofconstruction, and seallubrication play an impor-tant role in successfuloperation. Achieving ahigh level of reliabilityand service life is a clas-sic problem in the field oftribology, the study offriction, wear, and lubri-cation.

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

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

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

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

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

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

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

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

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

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

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

Q:elastically under contact pressure.This deformation creates largerfilm areas and very thin films.Such lubrication systems are nor-mally used to control wear inrolling element bearings. In seals

where the viscosity ofthe fluid sealed increas-es with increasing pres-sure, elastohydrodyna-mic lubrication occurs.

Boundary lubrica-tion is important forseal faces that are mov-ing very slowly underheavy load. Here,hydrodynamic and elas-tohydrodynamic lubri-cant pressures areinsufficient to separatethe seal faces. The slid-ing surfaces are protect-ed by the tribologicalproperties of the materi-als of construction. Anexample of a seal operat-ing within this lubrica-tion system is a dry-running agitator seal.

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

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

A

mechanical

seal is

designed

to operate

in many

types of

fluids.

A:


HANDBOOK??


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

five elements that can be quanti-fied and evaluated:

1. Temperature rise (minimumthermal flow)

2. Minimum stable flow

3. Thrust capacity

4. NPSH requirements

5. Recirculation

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

TEMPERATURE RISETemperature rise comes from

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

H 1∆T = ————— x ——————

778 x Cp η – 1

where

H = total head in feet

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

η = pump efficiency in decimalform

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

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

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

mechanical handbooks.What is the maximum allowable

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

allowable temperature. The differ-ence between the allowable tem-perature and the temperature at thepump inlet is the maximum allow-able temperature rise. Knowing ∆Tand Cp, 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

FIGURE 1


pump in Figure 1. Although pres-sure increases as the fluid ispumped upward through thestages, consider the pump inlet.The fluid at the inlet (low pres-sure, low temperature) is exposedto the temperature of the fluid inthe discharge riser in the head(higher pressure, higher tempera-ture). This means that the vaporpressure of the fluid at the pumpinlet must be high enough toaccommodate the total tempera-ture rise through all the stages. Ifthis condition is discovered duringthe pump design phase, a thermalbarrier can be designed to reducethe temperature that the inlet fluidis exposed to.

Some books, such as thePump Handbook (Ref. 5), contain atypical chart based on water (Cp= 1.0) that can be used with themanufacturer’s performance curveto determine temperature rise. Ifthe maximum allowable tempera-ture rise exceeds the previouslydetermined allowable temperaturerise, a heat shield can be designedand fitted to the pump during thedesign stage. This requirementmust be recognized during thedesign stage, because once thepump is built, options for retro-fitting the pump with a heat shieldare greatly reduced.

MINIMUM STABLE FLOWMinimum stable flow can be

defined as the flow correspondingto the head that equals shutoffhead. In other words, outside the”droop“ in the head capacitycurve. In general, pumps with aspecific speed less than 1,000 thatare designed for optimum efficien-cy have a drooping curve. Gettingrid of this ”hump“ requires animpeller redesign; however, notethat there will be a loss of efficien-cy and an increase in NPSHR.

What’s wrong with a droop-ing head/capacity curve? A droop-ing curve has corresponding headsfor two different flows. The pumpreacts to the system requirements,and there are two flows where thepump can meet the systemrequirements. As a result, it”hunts“ or ”shuttles“ between

these two flows. This candamage the pump andother equipment, but it willhappen only under certaincircumstances:

1. The liquid pumpedmust be uninhibited atboth the suction anddischarge vessels.

2. One element in thesystem must be able tostore and return ener-gy, i.e., a water col-umn or trapped gas.

3. Something must upsetthe system to make itstart hunting, i.e., start-ing another pump inparallel or throttling avalve.

Note: All of these mustbe present at the same timeto cause the pump to hunt.

Minimum flow basedon the shape of the perfor-mance curve is not somuch a function of thepump as it is a function ofthe system where thepump is placed. A pump ina system where the abovecriteria are present shouldnot have a drooping curvein the zone of operation.

Because pumps with adrooping head/capacitycurve have higher efficiencyand a lower operating cost, it wouldseem prudent to investigate the instal-lation 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. These

bearings can be sized to handlethe thrust. Thrust can be balancedby the use of balanced and unbal-anced stages or adding a balancedrum, if necessary. These tech-niques for thrust balancing areused when high thrust motors arenot available. It is worth notingthat balanced stages incorporatewear rings and balance holes toachieve lower thrust; therefore, aslight reduction in pump efficien-cy can be expected, and energycosts become a factor.

NPSH REQUIREMENTSHow many pumps have been

oversized because of NPSH avail-able (NPSHA)? It seems the easiestsolution to an NPSH problem is togo to the next size pump with a

FIGURE 2

Recirculation zones are always on thepressure side of the vane. A shows dis-charge recirculation (the front shroudhas been left out for clarity). B showsinlet recirculation.

A

B


this is the case, why don’t morepumps have problems?

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 the

larger suction, thereby reducingthe inlet losses. A couple of factorsbecome entangled when this isdone. A larger pump means oper-ating back on the pump curve.Minimum flow must be consid-ered. Is the curve stable? Whatabout temperature rise? If there isalready an NPSH problem, anextra few degrees of temperaturerise will not help the situation.The thrust and eye diameter willincrease, possibly causing damag-ing recirculation. When trying tosolve an NPSH problem, don’ttake the easiest way out. Look atother options that may solve along-term problem and reduceoperating costs.

RECIRCULATION

Every pump has a pointwhere recirculation begins. But if

10 15 20 25 30 35 40

vane at the outlet (Figure 2). Thesolution to this problem lies inthe impeller design. The problemis the result of a mismatched caseand impeller, too little vane over-lap in the impeller design, ortrimming the impeller below theminimum diameter for which itwas designed.

Recirculation is one of themost difficult problems to under-stand and document. Many stud-ies on the topic have been doneover the years. Mr. Fraser’spaper (Ref. 1) is one of the mostuseful tools for determiningwhere recirculation begins. In ithe describes how to calculate theinception of recirculation basedon specific design characteristicsof the impeller and he includescharts that can be used with aminimum amount of information.An example of Fraser calcula-tions, which show the require-ments to calculate the inceptionof suction and discharge recircu-lation, is shown in Figure 3.

RECIRCULATION CALCULATIONS

Figure 3 indicates the user-defined variables and chartsrequired to make the Fraser calcu-lations for minimum flow.Information to do the detailed cal-culations include:

Q = capacity at the best efficiency point

H = head at the best efficiencypoint

NPSHR = net positive suctionhead required at the pumpsuction

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

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

FIGURE 3

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

Cm2U2

Discharge Angle β2 Inlet Angle β1

VeU1

.14

.12

.10

.08

.06

.04

10 15 20 25 30.02

.10

.12

.14

.16

.18

.20

.22

.24

.26

.28

.30

.32

R1

R2

D2

D1B1

B2

h1

.08


The above information isobtained from the pump manufac-turer curves or impeller designfiles. The impeller design valuesare usually considered proprietaryinformation.

KVe and KCm2 can be deter-mined from the charts in Figure 3.

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 accurate-

ly determined if the elementsdescribed above are reviewed bythe user and the manufacturer.Neither has all the information todetermine a minimum flow that iseconomical, efficient, and insures atrouble-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 centrifugal

pumps. Chemical Engineering(1984).

4. I. Taylor. NPSH still pumpapplication problem. The Oiland Gas Journal (1978).

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

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

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

BY JIM MILLER

Many times I am askedto predict the flow of apositive displacementpump operating underconditions with insuffi-

cient inlet pressure to ensurecomplete filling. In these circum-stances the mixture may be con-sidered two-phase, with eitherfluid vapor or some otherentrained gas (air) present. Isthere any literature dealing withthe theory necessary to evaluatethe filling capability?

Adequate suction pressurefor a reciprocating positivedisplacement pump is anoften misunderstood con-cept. Pump manufacturers

are asked the minimum suction pres-sure that can be used for both contin-uous and intermittent service for aparticular pump. The net positivesuction head (NPSH) required may ormay not be provided. NPSH requiredtakes into account the fluid propertiesof water (usually), pump dynamics,and pump design. When analyzing apump’s operation at very low suctionpressure, several pump parametersmust be studied.

NPSH REQUIREDThe subject of NPSH for such

pumps is also frequently misunder-stood. The definition in the HydraulicInstitute Standards of minimumNPSH required is established for aspecific pump, pump speed, and dis-charge pressure by reducing the suc-tion pressure until one of thefollowing conditions occur:

1. a 3% drop in capacity

2. clearly audible cavitation noise

Several things are wrong withoperating a pump at the NPSHrequired suction pressure. By defini-tion, the pump is cavitating whencondition 2 occurs. Loss of capacityaccording to condition 1 means that

Q: the liquid chamber is only partiallyfilling as the result of cavitation.

NPSH required accounts for1) Fluid properties

a) vapor pressures

b) gas saturation pressure

c) fluid compressibility

2) Mechanical design

a) chamber volume

b) valve design

c) piping system

3) Operating conditionsa) pump speed

b) discharge pressure

c) fluid temperature

The meaningfulness of the NPSHrequired will become questionable ifany of these parameters change sig-nificantly for a given application.

PUMP LIQUID CHAMBERIt is difficult to predict whether a

pump will operate with some degreeof vapor or gas breakout. Given aknown volume and pressure of thegas or vapor escape, the solution isrelatively easy to solve. The difficultissue is the amount of gas initiallyformed in the pump chamber in thefirst place.

For the pump to continue tooperate, the cylinder pressure mustexceed the discharge manifold pres-sure at the end of the plunger’s dis-charge stroke. The followingapproach has been used to solve forpressure at the end of the dischargestroke:

Given:

∆volumeplunger = plunger displacement

∆volumeplunger =∆gas volume + ∆liquid volume

∆gas volume = v1 - v2p1= v1 - __ v1p2

p1v1= v1 - _______p1 + ∆p

∆liquid volume = βυ0∆p

∆volumeplunger =

p1v1= v1 - _______ + βυ0∆pp1 + ∆p

Solve for ∆p, the pressure

(βυ0)∆p2 + (βυ0p1 + v1 -∆volumeplunger) ∆p +(-∆volumeplungerp1) = 0

a = βυ0

b = βυ0p1 + v1 - ∆volumeplunger

c = -∆volumeplungerp1

-b ± √(b2 - 4ac)∆p = ________________2a

For a typical 2-in. plunger by 6-in.stroke pump the following valueshave been calculated where:

∆volumeplunger = 18.85 in.3

β = 0.000003/psi for water

liquid chamber volume = 100 in.3

υ0 = 100 (1 - gas fraction), (in.3)

v1 = 100 x gas fraction (in.3)

p1 = 30 psi

As can be seen in Figure 1, thispump operating at 1,000 psi dis-charge pressure will become vaporlocked when the gas volume is equalto 19% of the chamber volume. Thegraph suggests that a reciprocatingpump will operate at very low volu-

NPSH Required for Reciprocating PD Pump

A:


HANDBOOK??



metric efficiencies, but in fact it willstop pumping when it falls to the 75-85% volumetric efficiency range.This is probably because when thegas or vapor phase develops, it veryquickly exceeds the 19% volume ofthe liquid chamber.

CONCLUSIONSMinimum NPSH required suc-

tion pressure should not be used for acontinuous-duty pump application.As general practice, suction pressureshould be 10 psi above the minimumNPSH required for the pump. Thisadditional pressure will significantlyreduce the potential for pump cavita-tion problems such as damagedplungers, valves, and pump liquidends. The minimum NPSH requiredcould be used as a conservative suc-tion pressure to assure that the pumpwill continue to operate in an inter-mittent service.

Jim Miller is president of WhiteRock Engineering in Dallas, TX. He hasdegrees in chemical engineering andbusiness administration from theUniversity of Texas at Austin.

FIGURE 1

Maximum cylinder pressure versus liquid chamber gas fraction.

0 0.05 0.1 0.15 0.2 0.25 0.3Liquid Chamber Gas Fraction

100000

10000

1000

100

10

Max

imum

Cyl

inde

r Pre

ssur

e - p

sig


HANDBOOK??


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

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

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

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

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

RULE #1. PROVIDE SUFFICIENT NPSH

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

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

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

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

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

NPSHA = Ha + Hs - Hvp - Hf

where Ha= the head on the surface of the

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

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

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

Hf= the friction losses in thesuction piping.

The NPSH Available may alsobe determined with this equation:

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

whereHa= atmospheric pressure in

feet of head.Hg= the gauge pressure at the

suction flange in feet ofhead.

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

RULE #2. REDUCE THE FRICTION LOSSES

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

RULE #3.NO ELBOWS ON THE SUCTION FLANGE

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

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

Pump Suction ConditionsBY ROSS C. MACKAY

I

2g

FIGURE 1

Ha

Hvp

Hf

Hs


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 suction

of the pump with astraight run of pipein a length equiva-lent to 5-10 timesthe diameter of thatpipe. The smallermultiplier would beused on the largerpipe diameters andvice versa.

RULE #4. STOP AIROR VAPOR ENTERINGTHE SUCTION LINE

Any high spotin the suction linecan become filled with air or vaporwhich, if transported into theimpeller, will create an effect simi-lar to cavitation and with the sameresults. Services which are particu-larly susceptible to this situation arethose where the pumpage containsa significant amount of entrainedair or vapor, as well as those oper-ating on a suction lift, where it canalso cause the pump to lose itsprime. (Figure 3)

A similar effect can becaused by a concentricreducer. The suction of a

pump shouldbe fitted withan eccentricreducer posi-tioned withthe flat sideuppermos t .(Figure 4).

If a pumpis taking itssuction from a sumpor tank, the formationof vortices can drawair into the suctionline. This can usuallybe prevented by pro-viding sufficient sub-mergence 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 worthwhile noting thatthese vortices are more difficultto troubleshoot in a closed tanksimply because they can’t beseen 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 mayrequire 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 sim-ply passed throughto the pump.

RULE #6. WHEN RULES 1 TO 5HAVE BEEN IGNORED,FOLLOW RULES 1 TO5.

Piping design isone area where thebasic principles in-volved are regularlyignored, resulting in

hydraulic instabilities in theimpeller which translate into addi-tional shaft loading, higher vibra-tion levels and premature failure ofthe seal or bearings. Because thereare many other reasons why pumpscould vibrate, and why seals andbearings fail, the trouble is rarelytraced to incorrect piping.

It has been argued that becausemany pumps are piped incorrectlyand most of them are operatingquite satisfactorily, piping procedureis not important. Unfortunately, sat-isfactory operation is a relative term,and what may be acceptable in one

plant may be inappropriate in anoth-er.

Even when ”satisfactory” pumpoperation is obtained, that doesn’tautomatically make a questionablepiping practice correct. It merelymakes it lucky.

The suction side of a pump ismuch more important than the pip-ing on the discharge. If any mis-takes are made on the dischargeside, they can usually be compen-sated for by increasing the perfor-mance capability from the pump.Problems on the suction side, how-ever, can be the source of ongoingand expensive difficulties whichmay never be traced back to thatarea.

In other words, if yourreceivers aren’t performing well, isit their fault? Or does the quarter-back need more training? ■

Ross C. Mackay is an independentconsultant who specializes in advancedtechnology training for pump mainte-nance cost reduction. He also serves onthe editorial advisory board for Pumpsand Systems.

Inflow Inflow

To PumpSuction

To PumpSuction

Baffle

FIGURE 5


HANDBOOK?


BY ROBERT KREBS, CONTRIBUTING EDITOR

professional engineer came tothe Pumps and Systems boothat the Water EnvironmentFederation Exposition inAnaheim last October. He

asked to talk to someone regarding apumping problem. I happened to bethere and volunteered. Here is hisstory.

Two pumps are installed, asshown in the Figure 1 sketch. Thepumps are taking suction from a con-stant level water source. There isminimal velocity (much less than 1ft/sec) in the stream flowing by thepump suction line. The foot valve inthe suction line prevents flow fromthe pump to the source on pumpshutdown and on initial primingwhen the suction line from thepumps to the source are manuallyfilled with liquid. The pumps alter-nated in service and the ”off“ pumplost prime and would not start whenactivated. The engineer surmised thatthe ”off“ pump’s suction line wasbeing partially dewatered by the ”on“pump and would not prime on thenext start. The check valves in thesuction lines were added to the sys-tem. Periodically, the pumps still lostprime and would not function.

The application sketch (Figure 1)was prepared and discussed with theengineer. The suction line size andthe pump selection were satisfactoryfor the flow and pressure required.The system had been checked forsuction side leaks. The shaft sealingsystem was so designed that air wasnot entering the pump casingthrough the stuffing box.

In reviewing the installation as itwas originally designed (before checkvalves were added in the suctionpipes), I figured that the foot valvewas probably slowly leaking and/orthe reduction in suction pressure atthe ”off“ pump was bringing air out ofsolution at the lowest pressure loca-tion—the ”off“ pump casing. The pres-sure at either pump suction wasalways below atmospheric pressure.During shutdown, air could form inthe casing. On startup, there was not

sufficient liquid in the casingto prime the pump. Thewater being pumped is theeffluent from a waste watertreatment plant. The finalprocesses in these systemssaturate the water with air.Entrained air reacts to evenslight pressure reductions bycoming out of solution andinto a gaseous phase.

Adding the suction sidecheck valves to the systemtrapped liquid in the suctionline to the pump and stabi-lized the ”off“ pump suctionpressure, assuming thecheck valves did not leak.

The fact that the ”off“pump would periodicallystill lose prime when startedhints that the foot valve is leakingsufficiently to reduce the liquid vol-ume in the suction line enough toprevent pump priming.

Single suction lines for eachpump from an atmospheric pressuresource are recommended for allinstallations – flooded suction or suc-tion lift. On applications from aprocess vessel, a single suction linefor two pumps must be designed toassure uniform flow to either orboth pumps. Solid bearing liquidsrequire special consideration.

Foot valves are not recommend-ed as a conservative design practicein an unattended application of acritical pumping task because inter-mittent or continuing leakage is apotential risk. Foot valves of alldesigns have high friction losswhich increases suction side losses.Any solids in the liquid increase theprobability of leakage, and footvalves should not be used in solidsbearing liquids.

Suction lift system designsinvolving liquids with high vaporpressures or high levels of dis-solved air should be limited toshort lifts depending on individualconditions. In another column wewill discuss calculating net positivesuction head available (NPSHA) on

Where’s the Prime?

A

liquids with high levels of flashinggases or dissolved air. Pump designsfor suction lift applications shouldinclude stuffing box shaft sealing sothat atmospheric air is not drawninto the pump.

How can the problem in thisinstallation be rectified ... what isthe most cost effective solution...

FIGURE 1

Chlorination Water Pump System

Pump 1

Pump 2

To System

CheckValves

Foot Valve

5 ft.

10 ft.

FIGURE 2

Centrifugal Pump Priming

Foot Valve

PrimingChamber

Fill Port

Low LiquidLevel


and how can the problem be avoid-ed in a new installation?

Retrofit methods and designoptions can include: 1. separate suction lines to each

pump with foot valves 2. installation of a priming chamber

with the existing piping

3. utilization of submersible pumpslocated in the liquid sump

4. installation of an automatic prim-ing system

5. use of a self-priming end suctioncentrifugal pump

A reasonable assumption is thatthe existing foot valve is leaking nowor will leak in the future. Installing apriming chamber could be the mostcost effective solution (Figure 2). Thesuction line is raised to the top of thepriming chamber which is initiallyfilled through the fill port. On shut-down, should leakage occur throughthe foot valve the priming chamberliquid level will drop no lower thanthe suction inlet.

The priming chamber volumemust be several times (3-5) that of thesuction pipe volume. When the pumpstarts, the liquid level in the primingchamber drops, and the pressure inthe priming chamber will fall belowatmospheric pressure. As pumpingcontinues, atmospheric pressure onthe source surface will force liquid upthe suction pipe and into the primingchamber assuring continued pump-ing. This would be a positive fix to theexisting installation.

For the engineer’s existing situa-tion, the less costly solution would beseparate suction lines with non-metal-lic ball check valves and seats as thefoot valves. The pumps should againbe checked out to be sure the shaftsealing method is not allowing air toenter the pump. If shaft packing isused, a water sealed lantern or seal ringmay be necessary, and seal waterwould have to run continuously.

If the foot valves leak, a simplevacuum priming system might beadded. If pressurized water is avail-able, an ejector priming system couldbe a minimal cost addition. Compressedair or steam can be substituted forwater in the ejector priming system.One form of an ejector priming sys-tem is shown in Figure 3. Other formsof priming systems use vacuum pumpsto reduce pressure and remove air inthe pump and suction lines.

If this application is new, thedesigner might consider submersiblepumps, self-priming pumps, or endsuction centrifugals with an automat-ic priming system. The pump choicewould be set by the flow and pres-sure required. ■

Automatic Priming System

FIGURE 3

ControlValve

PressurizedFluid

Liquid Sensor

To Waste

Venturi Ejector

Vent Air Valve


HANDBOOK??



plant engineer for a Southerncooperative utility sent us aletter recently expressing theconcerns we hear so manytimes from individuals who

have the occasional need to workwith pumps and systems. Each timethese people get involved withpumps, they must retain systemdesign, pump types and sizes, frictionloss calculation, and piping design.

The five questions in his letteraddressed these concerns and askedfor sources to help the infrequentuser. To address this need, we willbegin, with this issue, a series ofcolumns that will start with simplesystem designs and move toward themore complicated configurations andpump selection.

Although some of this may be”old hat“ to the more experiencedreader, there may be others in yourorganization who could profit fromthis information, so pass it along tothem.

Let’s take a trip through a pumpsystem. First, there has to be a needfor a pump—a need to raise the ener-gy level of a liquid to accomplish apurpose. A pump system takes liquidfrom a source and delivers it to a dis-charge point. It consists then of asource and destination point, valves,a pump or pumps with drivers andpiping as shown in Figure 1.

There are friction losses in thepiping (suction and discharge), thepipe fittings and the valves whichincrease with flow rate. Tables of fric-tion loss are available in hydraulichandbooks and in the engineeringsection of vendor catalogs.

The pump requires a certainminimum pressure at its entrance orsuction point. This minimum pres-sure increases with flow and pumpspeed. This is called the net positivesuction head (pressure) required(NPSHR) above the vapor pressure ofthe liquid. Therefore, the systemmust provide at least that net positivesuction head available (NPSHA)above the vapor pressure of the liq-uid at the design flow conditions.

A All pump sys-tems can be catego-rized by liquid des-tination as eitheropen-transfer orclosed-circulatingsystems. Figure 1 isan open transfer sys-tem. The source anddestination may beopen to atmosphereor closed and sub-ject to different pres-sures. Figure 2 is aclosed-circulating system. The sourceand destination are the same. AtStation A, work is performed on orby the liquid, and the liquid isreturned to the source. Thepump is supplying the energy tocirculate the liquid through thesystem. An example would be aheat transfer system.

Both open-transfer andclosed-circulating systems mayhave multiple destinations as inFigures 3 and 4.

The destinations examples inFigure 3 may all be at differentpressures and flow rates, requir-ing pressure regulation and flowcontrol at each destination. Toprovide for continuous pumpoperation a control valve bypassto the source may be necessary.Other options to achieve continu-ous operation are variable speeddrive and differential system pressureoperation.

In the closed cir-culating system, Fig-ure 4, the destin-ations may also benon-uniform as toflow and pressureand could requirepumping to return tothe source.

Multiple pumpstations may beincluded in open orclosed pumping sys-tems (Figures 5 and6). Figure 5 is a mul-tiple pump single

Pump System Design - Part 1

FIGURE 1

FIGURE 2

FIGURE 3

A


destination open system. Both pumpstations (PS) 1 and 2 are designed todeliver liquid to destination A.

In Figure 6, PS1 is providing flowto both A and B. PS2 could bedesigned to deliver to A and PS3 todeliver to B.

It is obvious that the arrange-ment of pump systems is virtuallyunlimited. However, the analysis ofalmost all types can be simplified tosome form of these six examples.

The system is analyzed to deter-mine the proper location and require-ments for the pump. The systemdesign will provide the pump suppli-er with the criteria needed for theapplication. The pump supplier willrespond with an offering.

Designing the system is illustrat-ed by taking an open-transfer systemand working through the example.The same approach would be usedon a closed-circulating system. Thedesigner gathers data on the liquidproperties, source and destinationlocation, pipe routing and lengths, the need for special regulatory con-siderations, and flow rates. A flowsketch, as in Figure 7, is made andthe data for design entered. Thedesigner then constructs a hydraulicflow picture of the system resistancecalled a system head curve.

Figure 7 is an open-transfer sys-tem with the system resistance curve.The source pressure (Ps) and suctionstatic head (hsu) are resisted by thefrictional pipe resistance of the suc-tion side (hfs). The pump adds suffi-cient energy to the liquid at thedesign flow rate to overcome the sumof the static head (h-hsu) and pressuredifferential (Pd-Ps) and the total pipefriction loss. The pump produces adifferential pressure called the totalhead H1 at the design flow rate Q1.

For this transfer system, thetotal head H1 can be expressed as

H1=discharge side pressure –suction side pressure +friction losses

H1=[Pd+h] – [Ps+hsu] + [hfs+hfd]

Note that the pump produces adifferential pressure—discharge pres-sure at the pump discharge minus thepressure at the pump suction. Thesuction pressure must be calculated

˙ FIGURE 4

FIGURE 5

A

2

1

FIGURE 6

B

A

2

1 3


to determine that adequate NPSHA is available.In this example, suction pressure would beequal to Ps+hsu – hfs.

The total head will vary with pipe size andwith progressive fouling of the interior wall ofthe pipe in use. Since the pump can only operateat the intersection of its performance curve withthe actual system head curve, determination ofan accurate system resistance is key to a gooddesign.

Example 1 adds numbers to Figure 7. Thesolution will be discussed next month. Get outyour handbook and give it a try. The solutionwill be based on the Darcy-Weisbach pipe fric-tion formula and minor loss coefficients. Youmay use Chapter 3 of the Cameron HydraulicData book or the Engineering Data Book fromthe Hydraulic Institute (Table III B-4, lst Edition)that has almost the same values.

Example 1 Substitute the information belowinto Figure 7. Select a pipe size(s) and construct asystem head curve.

Pipe schedule 40 steel (new). Liquidpumped water 70 F atmospheric pressure 14.7psia. Design flow 200 gpm.

Ps = 10 PSIG Pd = 30 PSIG Pipe lengths-100 ft suction-1000 ft discharge Fittings as shown-std radius els-swing check& ball valves h = 100, hsu 20, hst 80. ■

FIGURE 7

Ps

Pd

hfshfd

hsu

hst

h

Smaller PipeSystem Head

Larger Pipe

hfs + hfd

hst + Pd - Ps

Total HeadH1

H

Q1Q


HANDBOOK??



ast month we discussed pumpsystem design. We also pro-vided an example of a singlepump, single destination

transfer system for you to solve forpump performance requirements.Table 1 gives friction loss data forthree pipe sizes and minor loss coeffi-cients for fittings in that pipe sizerange. The calculation, Figure 3, sum-marizes the 4-inch pipe results withthe system head curve drawn for 3-and 4-inch pipe.

Reviewing the calculation, theminor loss coefficients (K) wereadded together for the suction anddischarge sides. The table was con-structed to tabulate the total suctionand discharge side friction losses asa function of flow (gpm). Frictionfactor values per 100 ft and thevelocity head values were then mul-tiplied by the pipe lengths and totalK values, respectively. The plottedsystem head curves for 3- and 4-inch pipe are the calculated totallosses added to the static head (hsu).

If this pump system weredesigned for 200 gpm with 4-inchpipe on suction and discharge, thetotal head would be 110.1 ft or, with3-inch pipe, 183.7 ft. Althoughrarely used in water service, 3.5-inch steel pipe is available. The suc-tion side losses should be carefullyexamined. In this example, the staticsuction head (hsu) is 20 ft and thesource pressure (PS) is 20 psig(46 ft). The pressure availableat pump suction then is 66 ft,less the friction loss on the suc-tion side or 55.8 ft for 3-inchpipe and 63.4 ft for 4-inchpipe. The NPSHA should beadequate.

The optimum pipe sizewill take into consideration theinstalled cost of the pipe(which increases with pipediameter) and pump power(which increases withincreased friction in smallerdiameter pipe). A reasonableplan to start with would limitfriction loss at design flow to 2-

L


Steel Pipe Schedule 40 Minor Loss Factor K for FittingsFriction ft/100 ft/Velocity Hd (V2/2g) ft (Avg. for 3"-6") h(ft)=K x V2/2gFlowgpm 3 inch 4 inch 6 inch Description K50 0.7/0.1 0.2/0.02 0.03/0.01 45 el 0.27100 2.4/0.3 0.6/0.1 0.09/0.02 90 el 0.51150 5.1/0.6 1.3/0.2 0.2/0.05 Ball Valve 0.05200 8.9/1.2 2.2/0.4 0.3/0.08 Check Valve 1.7250 14/1.8 3.5/0.6 0.4/0.1 TEE Through 0.34300 20/2.6 4.9/0.9 0.6/0.2 TEE Branch 1.02400 34/4.7 8.5/1.6 1.1/0.3 Entrance Loss 0.05

Exit Loss 1.0

TABLE 1. PIPE FRICTION LOSS FOR WATER (APPROXIMATE)

System Conditions Flows: Destinations A, B, C, D, Each 100 gpm water (stp)Pressure: A, B, C, D, Each 40 psigSource: Open 0 psigPipes: 2,4,6 = 200 ft, 3,5 = 100 ft, 7 = 300-1,500 ft

Suction = 20 ft

FIGURE 1. SINGLE PUMP MULTI-DESTINATION SYSTEM

A B C

1

2

345

6

7

D

PS 1

10 ft

System ConditionsFlows: PS 1, PS 2, Each 200 gpm water (stp)Pressure: A = 32 psigSources: Open 0 psigPipes: 1 = 500 ft, 2 = 400 ft, 3 = 100 ft

Suction: PS 1 & 2 = 20 ft

FIGURE 2. MULTI-PUMP SINGLE DESTINATION SYSTEM

PS 1 PS 2A

2

31

10 ft

10 ft


5 ft of friction loss per 100 ft of pipe.High pipe velocities are not recom-mended on the pump suction side.Velocities of 5 - 10 ft/sec are suggest-ed. Discharge lines of longer lengthsare normally in the same range.Realize that sudden interruption offlow can result in large pressuresurges (water hammer) of about 50psi per ft/sec of interrupted flowvelocity on suction and discharge pip-ing.

In the example (Figure 3), 200gpm equates to 8.7 ft/sec in 3-inchpipe and 5 ft/sec in 4-inch pipe. Therespective friction losses are 9 and 2.3ft/100 ft of pipe. 4-inch pipe shouldbe used for this application. Thepump design would call for 200 gpmat 110 ft with 4-inch pipe.

We will go back to this examplewhen we start pump selection. Letme have your questions and com-ments. Nobody is perfect—no onehas all the answers.

The more complicated systemsrequire an initial approach analysis.Recall the analysis objective to estab-lish the performance characteristics inselecting a pump (or pumps) to meetthe pump system requirementsbecause the pump can only operate atthe intersection of the pump (orpumps) performance curve(s) and thesystem head curve.

Figure 1 shows a pump station(PS) with a multi-destination system.The pump and destinations A, B, Cand D are all at the same elevation.Destinations A, B, C and D would beequipped with flow and pressurecontrol to deliver the liquid ondemand. The maximum flowrequirement would be with all fourdestinations open and demandingtheir maximum flow. This would setthe capacity of pump 1. At flows lessthan maximum, a by-pass controlvalve in line 8 would open at somepressure above the set point and by-pass flow to the source. Alternately,the pump could operate on an on-offcycle if the frequency of cycling wasacceptable for the presumed electricmotor driver. Or a gas-pressurized(hydro-pneumatic) tank could besized to operate over a fixed pressuredifferential with intermittent pumpoperations. Still another option wouldbe variable speed control for the dri-

ver. However, with large and fre-quent variations in flow, variablespeed control may still require sys-tem storage or by-passing to avoidproblems.

How to decide on a pipe size?The easiest way for the experienceddesigner is to use one of the manyexpert pipe programs for personalcomputers. For the casual user (onceor twice a year), this may not be theanswer. Most computer programsare easy to understand when usedfrequently or after intensive trainingand use. Another way is to approxi-mate a result by designing for theworst condition, that is, the highest

flow and pressure requirements.Intermediate and low flow condi-tions affect pump selection, but notsystem design in clear liquid sys-tems. The worst condition scenariofor the system (in Figure 1) would bewith all four destinations at maxi-mum flow and pressure. However,there could be pump applicationproblems with this approach. Thosewill be covered later, when we dis-cuss pump selection.

Friction losses depend on flowand pipe size. In this example (Figure1), the longest pipe string is for flowto destination C through pipes 1, 3, 5and 7. With maximum flow to each

Minor Losses - Suction 90 el(0.51) Ball Valve(0.05) Entrance(0.5)Discharge 90 els([email protected]) Check Valve(1.7)Ball Valve(0.05) Exit(1.0)

Suction Side 4 inch Discharge Side 4 inch100 ft Total Total 1000 ft Total Total Total

Flow Pipe K(1.06) Suct. Pipe K(3.77) Disch. Friction Hd50 0.2 0.0 0.2 2 0.1 2.1 2.3100 0.6 0.1 0.7 6 0.4 6.4 7.1200 2.2 0.4 2.6 22 1.5 23.5 26.1300 4.9 1.0 5.9 49 3.4 52.4 58.3400 8.5 1.7 10.2 85 6.0 91.0 101.2

FIGURE 3. CALCULATIONS

3 in Pipe

System Head

4 in Pipe

Static Head

0 100 200 300 400Flow gpm

400

300

200

100

HeadFt


destination, the friction loss for eachpipe section can be manually calcu-lated and the pipe size adjustedwhere necessary. The pressurerequired at each destination plus thepipe friction back to the pump (withall destinations receiving liquid atmaximum flow) give pump dischargepressure for each destination. Thehighest of these pressure valueswould be the pump discharge pres-sure. Pump total head would be max-imum discharge pressure plusfriction loss to the maximum dis-charge destination less positive suc-tion pressure.

Figure 2 illustrates a multi-pumpstation, single destination, open trans-fer system. The general case would bewhere psi or pipe 2 or both coulddeliver flow up to some maximumvalue to location A. The complexity ofthe analysis of multi-pump stationproblems increases rapidly as thenumber of PS and destinations

increase. Again, the expert computersystems would be the solution choice.But, let’s examine a manual solutionmethod.

The objective is to determine theperformance requirements for psiand PS 2. With the required pressureknown at A along with the requiredflow from psi and PS 2, a pipe sizeand pressure drop for pipe segments2 and 3 can be calculated. The dis-charge pressure of PS 2 and the pres-sure at the intersection of pipesegments 2 and 3 can then be calcu-lated. This permits setting the pipesize and determining the friction lossin pipe 1 to determine psi dischargepressure.

This problem has, again, conve-niently assumed psi and PS 2, thepipe and location A at the same gradeelevation. Normally there are eleva-tion differences that must be factoredinto a solution.

A system head curve should bemade for all applications, where prac-tical. For either of the two cases(Figures 1 and 2), preparing a manualsystem head curve is complicated.Using one of the expert software sys-tems and varying the flow would pro-vide a range of performance for eachPS in a system. For single destina-tions and single PS systems, a spread-sheet template can be developed toproduce system head curves andpump(s) performance curve(s).

The system conditions shown forFigures 1 and 2 are to be used as twoproblem examples. Try workingthem out. We’ll give solutions nextmonth. Use the friction loss values inthe table. Let’s ignore the fittings.Getting the method right is moreimportant. ■

Until next time, when we will startpump selection, which will furtherdefine the system.


HANDBOOK??



ast month we demonstratedmore complex pump systemsand the methods to arrivemanually at pump perfor-mance requirements. Figures 1

and 2 (from March’s article; repro-duced showing solutions) illustratethe head loss for the pipe size indicat-ed at maximum flow conditions foreach pipe segment. For instance, inFigure 1 the maximum total head lossthe pump must produce would be thesum of the losses in pipe segments1,3,5 and 7, which is 19.8 ft. To thisamount add the 40 psig (92.4 ft) desti-nation pressure for a total dischargehead of 112.2 ft. The 10 ft static headminus the suction side head loss, 0.2ft, is subtracted from the total dis-charge head. The pump would berated at 400 gpm (100 gpm each forA,B,C,D) at 102.4 ft. This would bethe maximum pump requirement.

What is the performance require-ment if only one destination requiredflow? The lowest head requirementwould have destination A operational,with the balance of the 400 gpm flowrecirculating to the source. Flowthrough pipe segments 1 and 2 at 100gpm would have friction losses of 0.5ft and 4.8 ft, respectively. Whenadded to the static head at A (92.4 ft),the discharge head becomes 97.7 ft.Subtracting the net suction head (9.8ft), the required pump total head is87.9 ft versus 102.4 ft for all four des-tinations receiving flow. Each pumpmust be capable of producing 400gpm at 102.4 ft and operating to 87.9ft, where the flow will be greaterdepending on the design of the bypasssystem.

To summarize, the maximumand minimum pump performanceconditions have been determined.Since any other operational conditionwill fall between these values, apump suitable for the two conditionsshould perform in this application.Exact site conditions, periods of min-imum and maximum flow, variationsin destination pressure and flowwould also influence pump selection.

L Anticipating a question that mayarise—what about the pipe sizesselected? In choosing the pipe sizes,lower velocities are selected with rea-sonably uniform head losses formajor segments. With instantaneousinterruption of flow, surge pressures(water hammer) can occur under con-ditions present in this design. So thesystems should be checked, at leastby a rigid pipe analysis, for potential-ly dangerous surges.

The solution for Figure 2 showsthe head loss with both pump station1 (PS1) and PS2 operating. PS1 wouldbe rated at 200 gpm, plus the headloss in pipe segments 1 and 3, minusthe suction side static head plus thesuction line loss, for a total of 83.8 ft.Similarly, PS2 would be rated at 200gpm at 81.6 ft.

What are the conditions if onlyone pump station is operating? Line

segment 3 would have a lower fric-tion loss. Since each PS has 200 gpmflow if only one PS is operating, thefriction loss of pipe segment 3 wouldbe equal to 200 gpm flow or one halfthe flow with both PS operating.Since pipe friction is proportional tothe flow squared, the friction in pipesegment 3 with only one PS operatingwould be about one-fourth (actual is2.2 ft from the chart) of that with PS1and PS2 operating. This can becomean important factor in pump selec-tion. The pipe sizes are based on thesame reasoning as the previous exam-ple.

Figure 3 illustrates the solutionfrom Example 1 in the March article.The system head curves for 3-inchand 4-inch pipe are shown.Assuming a design flow of 300 gpm,the system head curve (static headplus pipe friction) shows a total head


6" – 0.2'PS 1

400 gpmD

A

1

10 ft.

345

2 6

7

B C

6" – 5.5'

3" – 4.8'

3" – 7.2'

3" – 4.8'3" – 4.8'

4" – 2.2'4" – 4

.9'

FIGURE 1. SINGLE PUMP MULTI-DESTINATION SYSTEM

PS 1200 gpm PS 2 - 200 gpm

A1

10 ft.

10 ft.

3

2

4" – 0.4'

4" – 0.4' 4" – 8.8'

4" – 8.5'4" – 11.0'

FIGURE 2. MULTI-PUMP SINGLE DESTINATION SYSTEM


of 138 ft. A pump must be selectedwith a performance curve that inter-sects the head and flow conditions.What type of pump?

Since the liquid is water—thepressure flow conditions are not rig-orous—we will assume the site condi-tions of the pump station permit ahorizontal electric motor drivendesign to be used. Take my word forit for now—a single stage end suctioncentrifugal pump will do the job. Ifyou are still interested, in the future Iwill present a generic method (usingvarious manufacturers) on how toselect a technically correct pump.

Figure 4 is a manufacturer’spublished performance curve for apump that satisfies the require-ment of 300 gpm and 138 ft. Whatcan we learn from this curve? Thepump is a 3x2x6.5 (that is, 3-inchsuction, 2-inch discharge and 6.5inches maximum diameterimpeller). The varying diameters(from 4.9 inches to 6.5 inches)reflect machining or trimming theouter diameter of the impeller.The operating speed for this pumpcurve is 3500 rpm. The hydraulicefficiency of the pump is 72% to87%, and the NPSH requiredvaries with flow rate from 4 ft to16 ft. The manufacturer has calcu-lated the brake horsepower andplotted lines of constant horsepow-er coincident with standard motorsizes of 5 hp through 15 hp. Thebrake horsepower required at anypoint on the performance curvecan be calculated by the formula:

BHP=QxH/3960xEff

For our condition (300 gpm at138 ft and 86% Eff from curve), thisequals 12.2 BHP.

What impeller diameter shouldwe select? Figure 4 shows the perfor-mance of a 6.5-inch and 6.1-inchdiameter impeller. The flow that willoccur appears at the intersection ofthese curves and the system curve.The flows are about 315 gpm and 275gpm for the 6.5-inch and 6.1-inchdiameter impellers, respectively. Asthe piping system continues, depositscan reduce the pipe’s internal diame-ter and will degrade the smooth sur-face of the new pipe. This aging

effect increases fric-tion loss. Thus,more head is re-quired to producethe same flow.

I would specifythe 6.5-inch impel-ler. How accurateis the calculatedhead? If all factorsare considered inthe system headanalysis, the calcu-lated head is proba-bly within 10% andis most probablyover the actualhead. What aboutthe source and des-tination conditions?Figure 4 assumesthem to be con-stant. That is rarelythe case. The cen-trifugal pump deliv-ers a differentialpressure. As thispressure reflectschanges at the source or destination,flow will change. The curve labelednew static head assumes a change insource/destination pressure. Now ifthe 6.5-inch diameter impeller is in-stalled, it will operate at approximate-ly 350 gpm under these newconditions. If the two static headconditions represent the maximumvariation that can occur in an appli-cation it is called the “operating headrange.”

The operating head range willvary with the number of pumps inservice at the pumping station. Thepump or pumps selected to operatemust be able to perform on the sys-tem head curve within the operatinghead range. The term “shut-off head”defines the pressure at 0 flow. It isimportant that the performance curverise uniformly to the shut-off head. ■

100 200 300 400

60

100

140

180

10 hp15 hp

7.5 hp

5 hp

4 8 12 16

7282 86

82

86

72

874.9"

5.6"6.1"

6.5"

NPSHR (ft)

Gallons per Minute

TotalHead (ft)

Pump 3x2x6.5 2000 Speed 3500 rpm

FIGURE 4: PUMP PERFORMANCE

FIGURE 3: GRAPH OF CALCULATIONS

400

300

200

100

0

HeadFt

3 in pipe

System Head

4 in pipe

Static Head

New Static Head

6.1 inch

6.5 inch impeller

Performance3x2x6.5 pump

0 100 200 300 400


HANDBOOK??



here is something wrongwith the pump.” Soundfamiliar? From my 40years of experience, I oftenfind that the pump is

blamed for a “system” problem. Yetmany, perhaps most, system prob-lems can be traced to suction sidesource or piping conditions. Thismonth, I will describe some recom-mended source and suction side pip-ing conditions, and analyze aninstallation.

Listed below are some desirablesource and suction piping conditionsfor shaft pump installation.1. The source should provide suffi-

cient pressure to provide therequired NPSH plus a margin of2 ft (8 ft for positive displace-ment pumps) at the pump suc-tion.

‘T 2. The source should be designedto avoid vortex formation and airor gas entrainment from turbu-lence.

3. The suction piping should bestraight, as short as possible, self-venting to the pump or thesource, or to waste, with a lowliquid velocity, preferably 5 ft/s.

4. If each pump cannot have a sep-arate suction pipe from thesource, the header piping and theimmediate pump suction piperequire special attention toensure minimum turbulence ofthe liquid stream into the pump.

5. Valves should be placed awayfrom the pump suction so thatflow to the pump is free from theturbulent effects they may cause.

6. Reducing fittings should be ofthe eccentric typemounted with thestraight side up.Steel and similarlydesigned reducingfittings shouldhave a straightsection before thepump suction.

Vertical shaft,submersible and othersubmerged pumpdesigns will havesome different re-quirements.

The objective ofthese six items is tofacilitate a designwhich avoids turbu-lence in the sourceand piping, and chan-nel the liquid to thepump suction with aslittle change in direc-tion or cross-sectionalarea as possible.

For example, aconsulting engineersent me a sketch of aninstallation (Figure 1)

with these questions: Will it work?What do you think?

The details of the installationshow one of two water system highservice vertical turbine pumps. Thepumps are taking suction from a clearwell of finished water with adequatesuction pressure through the 16-inchsuction line. Each pump is a 4-stagevertical turbine rated at 1800 gpm at240 ft head at 1180 rpm. The pumpsoperate at constant speed.

The pump suction bell is locatedat approximately the same elevationas the top of the 16-inch suction line.The velocity in the suction line isnearly 3 ft/s.

In the early 1960s, I served on aHydraulic Institute committee thatformulated some recommendationsfor sump designs and piping. Thesehave since been published in eachedition of the Hydraulic InstituteStandards. In analyzing this installa-tion, it is helpful to review the cur-rent (14th) edition, pages 126-133.

In analyzing any installation,there is a hesitancy to be critical ofdesign. Making general statementsabout a single situation is not recom-mended. However, when experienceshows that a design is suspect, as inthis case, it merits further study. It ispossible that this design would workfine, but I do not think so. Here is myanalysis.

With the system as shown, the 3ft/s velocity will interfere with the liq-uid making the 90-degree turn intothe suction in a uniform manner. Theresult is uneven filling or loading ofthe first stage impeller. The suctionflow pattern will possibly result invibration and premature wear of theimpeller.

To explain, consider the flowthrough a 90-degree elbow (Figure 2).The streamlines, which represent theliquid flow, are closer together at thelarger outer radius and further apartat the inner shorter radius. In theopen pump entrance, the same condi-tion will occur. More flow will tendto go to the outer or right side of thepump suction, as shown in Figure 1.

Pumping System Piping

FIGURE 1

Vertical turbine pump installation

16" Pipe

20" Pipe


How should this design be modified? If possible, thepump suction should be raised by 2 bowl diameters to allowthe flow to straighten. The suction pipe should be increasedto 20 or 24 inches. I favor an approach velocity (line velocityin this case) of about 1 ft/s with this design.

What can be done if the system is built and the problemsmentioned have occurred? I would suggest parallel turningvanes, as sketched in Figure 3, to help turn the liquid and dis-tribute uniform flow to the first stage impeller. ■

FIGURE 2

Flow through a bend

FIGURE 3

Turning vanes


HANDBOOK??



rom my experience, inertia oracceleration head is even lessunderstood among pumpusers than another hydraulictransient—surge or water

hammer. Inertia head should not beignored because it is too frequently acause of operating problems. Toexplain, let’s look at the example inFigure 1.

The centrifugal pumps are dri-ven by constant speed electricmotors. The control scheme calls fora pump to operate and a motor tostart. The motor and pump come upto full speed in a few seconds, andliquid flows through the dischargepipe. Right? Wrong? Maybe?

The 10,000 ft discharge line isfilled with liquid. This liquid massmust be accelerated from standstillto design velocity (3ft/s). ByNewton’s law (F=mA), this requiresan accelerating force ∆F = M(∆V/∆T) over a finite time period(∆T).

So what happens? When themotor starts, pump discharge pres-sure increases to meet system resis-tance. From Figure 2 theaccelerating force is the differencebetween the zero-flow head of thepump (Hso) and the static systemhead (Hs). This force is exerted as apressure on the liquid in the dis-charge line. The line velocity willgradually increase to the designpoint. The dotted line in Figure 2illustrates the immediate pressureresponse to the start-up signal. Theremay be a “pause time” at zero beforethe flow progresses to the designpoint—the intersection of the pumpcurve with the system head curve.

This phenomenon can beviewed at any pump station whereacceleration head is sufficient tocause a pause time. A pressuregauge on the pump discharge will,with pump start, indicate pressure inexcess of the anticipated operatingpressure, and after some time thepressure gauge reading will lower toan operating pressure.

FYou may be saying, “What’s the

big deal?” The answer is pump prob-lems. Centrifugal pumps are usuallyfurnished with an enclosed impellerand a volute-style casing. Whenoperated to the left of the best effi-ciency point capacity (lower flow),the liquid pressure in the casingvaries around the casing periphery,and the resultant unbalanced radialforce in the casing acts on the face ofthe impeller (between the shrouds)and deflects the shaft while increas-ing the radial bearing load. Since theforce is always in the same direc-tion, the shaft is experiencing areversed bending stress with eachrotation. If this stress exceeds theendurance limit of the shaft materi-al, a premature fatigue failure of theshaft will result.

Since the overstress conditiononly occurs on start-up, can it causedamage? Since stresses imposed inreversed bending are cumulative, itdepends on the number of starts, thepause time, and the amount of stressimposed. It should be mentionedthat loads imposed on the radial(normally inboard) bearing willreduce bearing life. Also, the deflect-ed shaft will increase packing leak-age or decrease mechanical seal life.

Fatigue failures in pump shaftscan also occur in rotary pumps

under some operating conditions,but that’s another story.

The approximate pause time canbe calculated from the formula:

∆T = (0.031) [LV / (Hso – Hs)]

which is derived directly fromNewton’s law. Figures 1 and 2 iden-tify the variables.

If a pump starts one time perhour, and the pause time is 10 sec-onds, and the pump rpm is 1800 (30Hz), then each day the pump willexperience 7200 revolutions of theshaft in the stressed mode. Mostfatigue failures occur in 1 millioncycles or less, or about 140 days ofoperation.

Figure 3 is one possible plot ofliquid line velocity from pump startto design velocity. This path func-tion can only be estimated frompoint to point with the formula.Obviously, pump flow will increaseto some value in excess of the manu-facturer’s limiting minimum flowprior to reaching design velocity.

Pump manufacturers place anoperating range of minimum andmaximum flow limit lines on theperformance curves to ensure satis-factory operation. Operating at lessthan the minimum recommendedflow can cause many problems.

Have You Broken Any Shafts Lately?

147 Ft Hs

L = 10000 FtV = 3 Ft/Sec

FIGURE 1. PUMPING SYSTEM


Inertia head problems are mostfrequently found in systems withsingle-stage, end- and double-suc-tion, overhung impeller centrifugalpump designs. Such designs oftenhave a high static head and lowcapacity. To pass large solids, non-clog impellers are designed with aminimum of 3-inch spacing be-tween the shrouds, providing alarge area for the unbalanced pres-sure to generate a large deflectingforce.

For many years I have used acalculated 10 seconds as the thresh-old pause time to require some sys-tem design change. Again, the designservice conditions must be consid-ered.

Other types of systems usingdiaphragm, blow case or pneumaticejector pumps should always bechecked for inertia head and suffi-cient air pressure accelerating force.

How to design around the prob-lem? For centrifugal pumps, a steepcurve with a high zero flow headwill maximize the available force forovercoming inertia. A solution is tobypass a portion of the dischargeflow to the source until the remain-ing residual force can start liquidflowing, at which time the bypasscloses. For clear liquid applicationsthere are special valve designs.Variable speed drives with infre-quent starts will also help. ■

Until next time. . . 3

2

1

05 10 Time

Sec

Pipe Design

VelocityFt/Sec

FIGURE 3. SYSTEM INERTIA EFFECT

Hso 240

Hs 147

HeadFt

Flow GPM

PauseTime

Recommended PumpOperating Range

Pump Perf. Curve

System Curve

Actual Pump performanceFrom Start-Up

FIGURE 2. PUMP & SYSTEM PERFORMANCE


HANDBOOK??



ometimes two pumps arerequired in the same service.This month we will look atsystem designs incorporatingtwo pumps which allow flow

contribution from one or bothpumps. In designs such as these, thepumps are said to be operating in par-allel.

In a simple case, two identicalpumps are selected to provide theflow (Figure 1 ). If they are centrifu-gal pumps, their performance on ahead-capacity chart will appear as inFigure 2a. The system head curvelocates the operating head and capac-ity for operating pump 1 or 2. Whentwo pumps operate in parallel, theircapacities at any head are additive,so the curve 1 + 2 shows the inter-section of the system head curvewith both pumps operational. Notethat the intersection with two pumpsoperating is at a higher head thanwith only one pump operating. Tofind the flow being pumped by eachpump with two operating, trace backat that head to the single pump curveand read the flow for each pump atPoint A. Since, with two pumpsoperating in parallel, the system headrises with increased flow, each pumpproduces less flow than when itoperates by itself.

What if the pumps are positivedisplacement? Figure 2billustrates such an example.The same rule applies—adding the flows of eachpump at the same pressureprovides the performanceof the two pumps operatingin parallel. Note the almostconstant flow rate withpressure. The slippage orbypass flow increases withpressure and decreases withincreased viscosity.

Pumps of different sizeoperating in parallel are alsocommon. Figure 3 illus-trates the system. If thepumps are centrifugal, theperformance will be asshown in Figure 3. If pump

S 1 or 2 is operated alone, the capacityeach would produce is the capacityshown at the intersection of thepump curve with the system curve.With both pumps operating, thecapacities are added at the same pres-sure. The resulting curve gives a newintersection point on the systemcurve for the total capacity.

To find the flow contribution ofeach pump, trace back at the headpressure to the intersection withpump curves 1 and 2 at point B anddetermine each pump’s performance.Care should be taken that the smallerpump is not forced to operate at aflow less than manufacturer’s recom-mendations.

If the pumps are positive dis-placement, their capacities are addedtogether at a given pressure. Theresulting intersection point with thesystem curve gives the total capacity.

In the case shown in Figures 2aand 3, the system head curve permitsboth pumps to contribute to the flowwhen both are operating. In somecases for centrifugal pumps, thesmaller pump or pumps may beunable to move any liquid with alarger pump operating because thesystem resistance exceeds the pumpshutoff head. Figure 3 illustrates thisconcept by changing pump 2 topump 3 (dotted line curve). Pump 1may be selected because it can pro-vide the flow for a high percentage ofrequirements with a lower horsepow-er than pump 3. Applying a variablespeed drive to pump 3 may accom-plish the same thing.

Figure 4 illustrates the impor-tance of insuring that the pumps willoperate individually in a system, aswell as in parallel. The two pumpsoperating intersect the system curve

Pumps in Parallel - When One Is Not Enough

2

3

1

FIGURE 1

FLOW FLOW

PUMP DESIGN

PUMP 1 OR 2

PUMP1 + 2

SYSTEM HEAD

A

PUMP 1 OR 2

PUMP1 + 2

PRES

SURE

HEAD

FIGURE 2

FIGURE A FIGURE B

2

1


within their flow capability.However, the system resistance atthe capacity limit of the identicalpump cannot be reached with onepump. Manufacturers place bothminimum and maximum capacitylimit lines on centrifugal models. Themaximum capacity limit or runoutcondition is usually 10-15% beyondBest Efficiency Point (BEP) capacity.Some pumps will pump more; how-ever the NPSHR for centrifugalpumps increases as a power function(to the 1.5-2.0 degree) with flow andmay exceed the NPSHA resulting incavitation.

Noise and vibration are otherhazards of operating at a runout con-dition.

Positive displacement pumpsused in parallel should have similarmaximum design pressures. Also, ifinternal and adjustable pressurerelief valves are present, they shouldbe set to correct pressures.

In these examples, we have con-veniently used identical suction anddischarge piping for each pump fromthe source to their common dischargeline. For instance, in Figure 1, ifpipes 1 and 2 are identical centrifugalpumps, the pressure at the dischargeof each pump would be the same.But what if pipe 1 is much smaller indiameter than pipe 2? The only pointof common pressure for the twopumps would then be at the intersec-tion of pipes 1 and 2 with the dis-charge line 3. What about the flow?The higher friction loss in pipe 1would meet the pressures at the 1-3intersection. The pump 1 dischargepressure would correspondinglyincrease, and the flow from pump 1would decrease.

What about the flow from pump2 under this condition? Note that thesystem head curve is independent ofthe pipe or pump. Note also thatpumps 1 and 2 operate independent-ly. The system curve is the control.Pump 2 continues to provide flow ata rate limited by the system resis-tance.

Some guidelines for pumps inparallel:• Use pumps in parallel as a back-

up in critical services.

• Use two or more pumps in paral-lel to meet fluctuating flowdemands or consider variablespeed control.

• Check how the pumps selectedwill operate in the system indi-vidually and/or in parallel if nec-essary to provide for adequatereserve.

• With several pumps operating inparallel, even if the largest pumpis out of service, the remainingunits should be able to pump themaximum flow.

• With identical pumps in parallelservice, use elapsed hour meterson each pump, along with alter-

nation, to assure that each pumpis regularly exercised to balancethe service life.

It should be clear that the accu-rate determination or calculation ofsystem resistance for the flow range isthe important step. ■

FIGURE 3

FLOW

B B

12

3

1 + 2

HEAD

FIGURE 4

FLOW

1 OR 21 + 2

MAXFLOW

HEAD


HANDBOOK??



hen matching a pump tothe system curve, one mayfind that the pump optionsavailable and suitable for

the service cannot meet the pressurerequirements at the design flow. Thedesign system pressure consists of sta-tic or elevation head plus the frictionhead or losses and any velocity headchanges at the design flow rate.

If the application requires a posi-tive displacement pump, the pumpmanufacturer’s literature should offervarying capacity pumps designed fora range of maximum pressures.However, meeting the design capaci-ty may require two or more pumps inparallel to achieve the desired flow.(See ”Pumps in Parallel” in theAugust issue.)

If the application calls for a cen-trifugal pump and a single stagepump will not produce the requiredpressure at the design flow, it is com-mon to use a two-stage or multi-stagepump, i.e., two or more impellersand casings on a single drive shaft,each producing the same flow to raisethe pressure a predeterminedamount.

For many services the searchends there. Centrifugal pump typesthat are manufactured in two- ormulti-stage designs include end suc-tion, double suction, vertical turbineand regenerative turbine pumps.

Some examples of designs avail-able only in single stage are non-clogdesigns for handling solids laden liq-uids, some slurry pump designs, andrubber lined pumps. For these pumpsthe pressure requirement may be metby employing two or more units inseries. (Crude oil and petroleumproducts pipeline pumps are highlyspecialized and not included in thisdiscussion.)

The August article on parallelpumping described how the flowrates of two operating pumps areadded at the same head or pressureto produce a composite performancecurve. Conversely, when two (ormore) centrifugal pumps are operatedin series within the same pump sta-

W

Pumps in Series - For More Pressure

Source (Constant Pressure) Destination

12

DesignPressure

HeadPressure

Design Flow

Flow

Pump 1 or 2

Pump 1 + 2

System Head

FIGURE 1. TWO CENTRIFUGAL PUMPS IN SERIES

Surge Tower

PS1PS2

H2

PS1

PS2

System Head PS2

Total Head PS1 = H1

Total Head PS1 = H1 + H2

System Head PS1

H1Design Flow

Flow

HeadPressure

DesignPressure

FIGURE 2. TWO CENTRIFUGAL PUMP STATIONS IN SERIES


tion, the heads or pressures producedby each pump are added at the sameflow rate to produce the compositeperformance curve.

The simplest configuration forcentrifugal pumps operated in seriesincorporates two identical pumpsconnected so the discharge of the firstenters the suction of the second,which then discharges to the system.Since the pumps are identical, eachproduces an equal pressure. Figure 1illustrates the performance of twoidentical pumps with the correlatingsystem head curves.

Pump two should be designedfor the higher suction and dischargepressures since the pressure at thestuffing box of this pump will behigher due to the higher suction pres-sure. The inter-connecting pipingshould be straight and no smaller indiameter than that of the pump suc-tion.

The series concept is complicatedfor pumps in separated pump stations(PS). Figure 2 illustrates a two PS inseries operation. The suction pressure

at the PS2 location must be carefullycontrolled, and PSl must provide ade-quate pressure at design flow todeliver liquid to the PS2 suction. PS2must then deliver the design flow tothe destination. Ideally, PS2 is locatedso that the total head of PS1 and PS2is the same. If PS2 is located somedistance from PS1, a surge tower orstandpipe at the second pump stationsuction might simplify the design (seeFigure 2).

Several years ago, I designed awater supply system that carried theraw water some 16 miles from thesource to a treatment plant located atan elevation more than 500 ft. abovethe source. To accommodate the lim-ited pipe design pressure, three boost-er pump stations were located so thattheir total heads were equal. Sincethere was no storage, water floweddirectly from one pump station dis-charge pipe directly into the suctionof the pumps in the next pump sta-tion. Calculations for surge pressureeffects and starting considerations

called for an array of control valvesand sensing systems.

If flow must move directly fromPS1 discharge to PS2 suction, thetotal head of each PS must be accu-rately calculated at the design flow.PS1 should be designed to providethe pressure for liquid to reach PS2with adequate suction pressure tomeet the PS2 NPSHR at the designflow. PS2 should be designed for itssystem requirements. As Figure 2illustrates, the problem becomes twoseparate applications.

These separated pump stationsconnected in series may produce amaintenance prone system. Toreduce the problem potential, thesame impeller design should beemployed for PS1 and PS2. Impellerdiameters and rotating speeds may bedifferent, but the curve shape willthen have the same characteristics.

If the application is suitable, theversatile vertical turbine pump designis the pump of choice for this typeapplication.


HANDBOOK?


BY TERRY M. WOLD

his article is an addendum toHerman Greutink’s articles“Vertical Turbine Pumps” and“Troubleshooting VerticalTurbine Pumps” (Pumps and

Systems, February 1994). All sugges-tions in these articles on installation,troubleshooting, and maintenance areapplicable to this discussion of verti-cal pumps with integral thrust bear-ings.

VERTICAL IN-LINE PUMPS WITH INTEGRAL THRUST BEARINGS

Vertical in-line pumps with inte-gral thrust bearings were originallydeveloped to facilitate in-line pumpmaintenance. In-line pumps have anAchilles’ heel when it comes to main-tenance. Since the driver bearings arealso the pump bearings, alignment ofthe pump and motor shafts is criticalto the life of the pump, especially themechanical seal and the close-run-ning clearances of throat bushing andwear rings. In-line pumps are usuallyof the volute design, and therefore aradial load is present. This results in ashaft deflection which is magnifiedby a combination of both final align-ment and machining tolerances. Theuse of a close-coupled design couldeliminate this difficulty, but thenthere are the complications from sealremoval, motor availability, producttemperature and limitation on sealingdevice types.

Vertical in-line pumps with inte-gral bearings are easier to maintainand thus have reduced maintenancecosts. Another use for these pumpsbecame apparent after initial usershad installed and operated a fewunits. Installation savings were real-ized because these pumps used a rela-tively small foundation (that is, lessspace) materials, and time for con-struction. The vertical in-line pumpwith integral thrust bearings could domost of what the traditional horizon-tal pump could do and save moneydoing it.

Vertical in-line pumps with inte-gral thrust bearings have these advan-tages:

T

1. The pump has its own bearingsfor handling thrust loads.Therefore, high thrust verticalmotors are not required, al-though a “P” base motor is rec-ommended due to tighter faceand runout tolerances, in addi-tion to the more uniform rabbetfit and bolt hole patterns.

2. A flexible disc type coupling canbe used instead of a solid axial-split spacer coupling, which canadd to alignment problems if notcorrectly manufactured, han-dled, and installed by trainedmechanics.

3. Since the radial bearing of thepump is closer to the impeller,shaft deflection is minimized andmechanical seal life is extended.

4. Foundation preparation and costare greatly reduced.

5. These pumps can be added to anexisting facility without massivepiping changes or relocation ofother equipment.

Factors that need to be evaluatedwhen considering the purchase of anin-line vertical pump with thrustbearings should include:

Vertical Pumps with Integral Thrust Bearings

Cross section of in-line pump with integral thrust bearings

FIGURE 1


• Bearing lubrication—oil mist oroil sump lubrication are recom-mended.

• Limitation of seal types—checkstuffing box space dimensionsand distance to the nearestobstruction.

• The rotating assembly should beremovable without disturbingthe pump case or driver. (This isan API 610 requirement.)

• Check the length of the couplingspacer. You should be able todisassemble the pump withoutremoving the driver.

• Check to see how the pump canbe handled during maintenanceprocedures. Consideration forlifting should be addressed.

VERTICAL TURBINE PUMPS WITHINTEGRAL THRUST BEARINGS

Vertical turbine pumps with inte-gral thrust bearings are becomingmore accepted in the United States.This is because only a few motormanufacturers will build a highthrust motor capable of handling theloads imposed by a multistage verti-cal turbine pump.

When evaluating the purchase ofa vertical turbine pump, you shouldconsider the following:

• For thrust balance considera-tions, the lower the thrust thelonger the bearing life. Minimumbearing life should be 25,000 to40,000 hours B-10, whether thebearings are in the driver or thepump. Some users have specified100,000 hours B-10 bearing life.To achieve this, the pump mustbe thrust balanced or a larger ordifferent type of thrust bearingmust be used. Special thrust bal-ancing is expensive. If the bear-ing size is increased, the ballsmay skid because of inadequatethrust load. If a different type ofbearing is installed, the lubrica-tion becomes complicated or themaximum allowable rpm isreduced. Although a higherpump efficiency can be attainedwithout thrust balanced im-pellers, the cost of maintenancedue to the higher thrust loads on

the bearings will exceed thesesavings.

• Pay close attention to the num-ber of alignment fits incorporatedin the pump design. The proba-bility of misalignment increaseswith more fits.

• Check how the adjustment forimpeller clearance is made. Isthere a one-piece shaft goingthrough the pump bearings,which is then connected by aflexible coupling to the driver?Or is there a rigid adjustable cou-pling below the bearings and ajack shaft and coupling above thebearings to couple the driver? Ingeneral, the one-piece shaft issuperior. The extra main-tenance to service the sealwhen one shaft is used faroutweighs the frequencyof seal repair due to theextra coupling and shaftalignment required whenusing a two-piece design.Those attending the lastAPI 610 8th Edition meet-ing agreed that the single-piece shaft is the prefer-red design. This sugges-tion may become part ofthe 8th edition.

• Check how maintenanceis performed on the bear-ing housing or mechanicalseal. The bearing housingis located between the dri-ver and the stuffing box.Therefore, removal of themechanical seal is per-formed by removing thebearing assembly first.You should be able toremove the seal and thebearing assembly withoutdisturbing the pump ormotor.

• Oil mist or oil sump arethe preferred methods oflubrication. In most casesan ISO 32 or ISO 68 non-detergent turbine oil issufficient.

• The recommended thrustbearing arrangement istwo 40° angular contactbearings arranged in a

back-to-back configuration withmachined bronze cages. With theexception of a momentary up-thrust at startup, it is preferredthat a pump has downthrustwhen operating at design condi-tion.

• Maximum oil temperature isdictated by the properties of theoil used. If the recommendedISO 32 or 68 oil is used, maxi-mum oil temperature should be 150°F. This can be greater if a higher viscosity oil isapplied, but two factors must bechecked. First, the minimumviscosity at operating tempera-ture should be 70 ssu. Second, ifa higher viscosity oil is used, the

Cross section of vertical turbine pumpwith integral thrust bearings

Motor Support

Coupling

Bearing Housing

Stuffing BoxPump Head

Column

Bowl

Impeller

Barrel

FIGURE 2


pumping characteristics of theoil circulation device will bedrastically impaired. Arcticapplications may require an oilheater or a lower viscosity oil,although a tropical environmentmay require a higher viscositylubricant.

• Vibration should be measured onthe bearing housing at the loca-tion of the bearings. It should notexceed 0.15 in/s between mini-mum flow and operating condi-tion.

• A conventional vertical turbinepump has a shorter overallheight and, with a properly sizedspacer coupling, the seal can beremoved without disturbing thepump or driver. However, align-ment becomes a more difficultand critical aspect.

CONCLUSIONVertical pumps with integral

thrust bearings have an expandingrole in industrial applications. Thistype of pump will not solve inherent

hydraulic problems but will increasethe mean time between failure inmost applications, as long as the cor-rect operating and maintenance pro-cedures are followed. These pumpsmay have a higher initial cost, butrepair savings may make them wellworth the investment. ■

Terry M. Wold is EngineeringManager at Afton Pumps, Houston,TX.

everal years ago I was con-fronted by the following situa-tion. The system, dia-grammed in Figure 1, was a

sewage pumping station with threeconstant speed centrifugal non-clogpumps of the same size. Pump 1would not pump; pump 2 wouldpump now and then; and pump 3worked fine.

As shown in the figure, therewere two inlets to the source wetwell. Source A was a gravity sewer,and pipe source B was a force maindestination from another constantspeed pumping station which periodi-cally delivered flow to the source wetwell at a rate of 550 gpm.

The first question we ex-ploredwas, why did pump 1 not pump?Was the impeller not turning ordamaged? Was there no water in thecasing? Was there a clog in theimpeller, or in the suction/dischargepipes or valves? Were there otherconditions we hadn’t considered thatcould be causing the trouble?

A visit to the station providedthese hints. Pump 1 would beginpumping, and then the flow would

gradually stop, as ifthere was no liquid inthe casing. Pump 2,after starting, wouldreact the same waywhen there was flowfrom the pipe Bsource. Finally, acrackling cavitation-like noise came fromall three pumps.

An inspection ofthe station and equip-ment ruled out clog-ging. The pumpswere new, so im-peller damage wasunlikely. The pumpshaft turned, and itwas doubtful that theimpeller was slippingon the shaft.

Other possibleculprits were groupedas follows:1. inadequate

NPSHA forpump NPSHR atthe design flow

2. air leakage to thecasing

3. operating totalhead much lessthan the designtotal head (i.e.,the pump notoperating on itscurve)

4. entrained air coming out of solu-tion and causing the pump to loseprime.

A second inspection demonstrat-ed that the minimum wet well levelwould ensure adequate NPSHA withrespect to NPSHR of the pumps atdesign flow. In addition, the pumpshad double mechanical seals whichwere lubricated with seal water, andthere were no apparent air leaks inthe pump station. Finally, a check oftotal head versus design head showed

that the pump was operating at a rea-sonable point on the curve.

Eliminating the first three possi-bilities left entrained air as the mostsuspect cause of the faulty pumpoperation. The discharge line fromthe upstream pumping station, pipeB, entered the station near the top ofthe wet well and was then directeddown into the well through an 8-in.cross fitting in a closed pipe to pre-vent excessive surface aeration in thewell. To avoid the 15 ft siphon effect,the designer provided a 1-in. openingin the cross cover plate as shown inFigure 2. However, no liquid came

Why doesn’t the pump pump?BY: J. ROBERT KREBS, P.E.

S

1

2

3

B

A

6" Suction

Sewer (elev. 938)

Pump Station

Concrete Filet

Wet Well

FIGURE 1


HANDBOOK??


��

��

Cap with "I" hole tappedin top – standard pipethread.

Cover plates

CL Elev. 950

8" cross

Float controls

Elev. 935

Bottom Elev. 933

Concretefilet

8" D.I.P.Force main

8' -0"wet well

FIGURE 2

from this opening when the upstreampump station operated.

When there was no flow fromthe upstream pump station, thewater level in pipe B was the same asthe level in the wet well. And, sincethe well liquid surface was at atmos-pheric pressure, the liquid level inpipe B was also at atmospheric pres-sure because of the 1-in. opening atthe top of the pipe.

However, when the upstreampump station delivered its rated flowof 550 gpm into the station throughpipe B, the pressure situationchanged. Since the pressure wasclearly atmospheric at the 1-in. open-ing, the pressure just inside the open-ing must have been less thanatmospheric to induce the no flowcondition.

What was happening? To answerthis question, we first analyzed the

forces acting on the liquid in pipe B.The 15 ft liquid column downwardvelocity is assisted by the gravityeffect of the elevation difference butresisted by the pipe friction. Theseforces balance at about 3000 gpm,and at lower flow the downwardgravity force dominates. Clearly,then, the pipe was not running full.And, because the down flowing liq-uid reduced the pressure in the verti-cal pipe B to below atmospheric, airentered the system through the 1-in.opening.

The end result was that a hugevolume of high velocity air enteredpipe B. This air left the pipe in closejuxtaposition to the pump 1 suctionpipe so that when pump 1 was oper-ating, substantial quantities ofentrained air would replace the liquidin the impeller and casing, causingthe pump to lose prime. Removing

the cross cover plate could reduce theturbulence of the air, but not thequantity of air.

The operating conditions of theupstream pump station were theninspected to confirm that the siphoneffect would not seriously reduce theoperating pressure of the pump in thatstation.

Closing the 1-in. opening on thecross provided the best solution. Thatwas done, and it worked. ■

Until next time . . .


90

80

95

85

120100A

BC

D

Elevations ft.

System Resistance

Static - Priming

Operating

Flow -gpm.Q

Headft.

+30

0-10


HANDBOOK??


hy pump downhill?Doesn’t water rundownhill naturally?While it’s true that in

an open to atmospheric pressuresystem (closed conduit or openchannel) the force of gravity caus-es the liquid to seek the system’slowest unrestricted elevationabove sea level, examples ofdownhill pumping systems are allaround us.

The pumping of effluent(treated waste water) into theocean through pipes that are fre-quently miles long provides onegood example of a downhill pump-ing application. Such ocean out-falls require pumping to overcomethe pipe friction loss at the desiredflow rate. Heating and cooling sys-tems in commercial buildings areanother example. These systemsare generally of a closed-circulat-ing type and the pumps are sizedto compensate for friction loss.

A more complex situation ispresented by raw water and wastewater pump stations. These sys-tems often require pumping fromthe source to a higher elevation fol-lowed by pumping to a lower eleva-tion destination. Many storm watersystems, for example, pump over ahigh point, such as a river levee,then discharge to a lower level, thuscreating an initial priming headand a lower operating head.

Let’s look at some essentialsof good system design for pump-ing downhill. Figure 1 illustrates acommon arrangement, a rawwater supply or waste water trans-fer system operating through anundulating terrain of peaks andvalleys. The highest elevation inthe system (point A) determinesthe static head which the pumpmust overcome at start-up.However, when the pipe is full,the static head will be a negativevalue based on the elevationsnoted in the figure. Theoretically,the siphon effect will supply flow,up to some value (Q), once the

pipe is filled. In fact, since the dis-charge destination, point D, is at alower elevation than the source forthe pump, liquid could (dependingon pump and control valve types)

continue to flow through the sys-tem after the pump stops. If thepump discharge pipe employs aswing or ball check valve, liquidcould flow through the stopped

Pumping DownhillBY J. ROBERT KREBS, P.E., CONTRIBUTING EDITOR

W

A

BC

Surge Relief Valve

Design Condition

Flow - gpm

Headft.

+30

0-10

VARVs

FIGURE 1. DOWNHILL PUMPING SYSTEM INITIAL CONCEPT

FIGURE 2. DOWNHILL PUMPING SYSTEM FINAL DESIGN


pump into the pipe, a conditionthat is to be avoided. An actuatorcontrolled plug or ball valve, sub-stituted for the check valve, willact as an isolation valve andensure that flow to the dischargepipe stops, as it should, with thepump.

Clearly air must be evacuatedfrom the pipe as it fills. Gooddesign dictates the use of properlysized air release valves (ARVs) atthe high points, A, B and C. TheARVs regulate the initial fillingrate of the discharge pipe and anysubsequent refilling of the pipeafter draining. The flow rate canbe controlled by throttling thepump discharge or reducingspeed, if variable speed is avail-able. The liquid velocity setpointwill vary with pipe diameter andthe steepness of slope (i.e., sharpor gradual) approaching the highelevation points. The objective inregulating the liquid velocity is topermit the trapped air in the pipeto leave as the pipe fills.

In addition, raw and wastewater systems generally containvarying amounts of entrained air.This air may come out of solutionand migrate to the high elevationpoints in the discharge pipe. TheARV selected for the applicationmust also be capable of removingthis air from the system.

With the ARV installed, the pipefilled and an open discharge at pointD, the control valve on the pump dis-charge will close when the pumpstops, prohibiting the water fromentering the pipe. Under these condi-tions, at the discharge point D waterwill flow via gravity to a lower eleva-tion, creating a lower pressure (orvacuum) in the pipe. Depending onthe elevation differences, a vapor cav-ity could form in the pipe, developinga pressure differential that could pos-sibly exceed the pipe’s threshold ofcollapse.

If the ARVs selected are combi-nation vacuum and air release valves(VARVs), these valves, at A, B, and C,will admit air from the atmospherewhen the pump stops. This partiallydrains the liquid at point D. TheVARVs must be sized to release air atthe high points with subsequentpump start-up and to allow air intothe pipe to relieve the vacuum forma-tion. As an alternative, a remotelycontrolled valve at point D could beclosed on pump shutdown.

At this point in the design, thesystem should be inspected for surgepressure. The system designer shouldbe aware that interrupting the flow inany system could cause the genera-tion of dangerous pressure surges orwater hammer. Provisions to attenu-ate these surge pressures for bothnormal and power failure shutdowns

are essential. System analysis forwater hammer is complicated andbest handled by specialists.

We are now ready to select apump for our downhill pumpingsystem. First, consider the operat-ing conditions. Figure 2 illustratesa specific system design with sys-tem head curves for both start-up/priming head and operatingconditions. A conventional swingcheck valve is employed at thepump discharge. Surge relief maybe provided by a controlled clos-ing check valve and/or a separatesurge relief valve. The points A, Band C are equipped with combina-tion vacuum and air release valves(VARVs). A constant speed pumpmay be selected if it can meet thefollowing criteria: • the ability to provide flow for

the initial pipe filling at the stationary priming head

• the ability to maintain thedesign flow at the operatinghead.

If these criteria can’t be met,a variable speed pump should beselected. ■



HANDBOOK??


magine you have just receiveda request to select a pump fora special project (and perhapsyou have). There are literally

hundreds of companies offeringone or more pump types to helpthe end-user solve that specialapplication problem. And, theoverwhelming number of choicesand questions may make selectinga pump a dizzying proposition.

Let’s try to sort a path throughthis maze.

Since the pump businessfocuses on application engineer-ing, it is imperative to begin with athorough systems analysis. Afaulty system design will certainlylead to improper pump selection–and poor operation.

Locating the pump in the sys-tem should be an early considera-tion. In addition, it is clearlyessential to determine how thepump will function in this system.Is the system functioning to trans-fer or circulate the fluid? Are thepump requirements continuous orintermittent?

The system analysis shouldstart with an evaluation of theprocess fluid. Does the solutioncontain solids? If so, what is thenature and concentration of thesolids? What is the viscosity, spe-cific gravity, vapor pressure andprocess temperature of the liquid?

Next, the analysis should con-sider the process variables, suchas: • net positive suction head

(NPSH) requirements

• the possibility of entrained airor gas

• source and destination pres-sure and temperature condi-tions and the anticipateddegree of variability in theseconditions

• flow rate and whether theflow is continuous, variable orintermittent.

The pump site and the environ-mental conditions for the project arealso important parameters in the sys-tem analysis. If the pump will beoperated in a clean, well ventilatedand protected area, the equipmentdemands will clearly be differentfrom those for a site out of doors invariable weather conditions. Waterbooster and waste water pump sta-tions, which may be located bynecessity in flood plain areas, alsoimpose unique site constraints.These stations often require specialpiping designs to avoid solids build-up, crystalline formations and/or toallow gas or vapor to escape.Additionally, determine whether thepump will be at a site that is heatedor cooled. Safety and hazard consid-erations associated with the locationshould be included in the analysis.Finally, weight and space limitationsas well as noise and vibration con-cerns must be taken into account.

Although system control andpump driver decisions are generallymade after the pump selection, thesite and environmental conditionsmay also affect these decisions. If so,these system control provisionsshould be included in the systemanalysis.

A completed design drawing willserve as an overall check of the sys-tem.

One may think of the pump interms of three separate design para-meters: hydraulic, mechanical andmaterials of construction. The sys-tem design will impose requirementsfor each of these parameters.

Now your job is to match thepump design to these requirements.This will assure a good workingpump and system. ■


Selecting a Pump-The Right Start Leads to the Right Finish

BY J. ROBERT KREBS, P.E., CONTRIBUTING EDITOR

I System and Pump Requirements

Environmental Considerations• location – temporary or permanent• hazards – electrical, mechanical or

vapor• site – flooding or equipment protection

Liquid Properties• chemical name• percentage of solids- their size

range• viscosity• specific gravity• temperatures• vapor pressure• pH

Hydraulic Design• capacity (min/max)• pressure-discharge and suction• static head-suction and discharge• NPSHA above vapor pressures• design (rated) capacity and differen-

tial pressure

Mechanical Design• expected mean time between failures• L-10 bearing life hours• intermittent or continuous service• variable or constant speed• casing design pressure• hydrostatic test pressure

Materials Design• consideration for abrasive or erosive

wear• corrosion allowance• failure hazard concerns

Drive and Control• electric motor or engine• air or hydraulic power• variable or constant speed• control of flow or pressure• valves, sensors and actuators



HANDBOOK??

he pumping system designis completed. Now thepump selection processbegins. The principal pump

operating criteria evolves from thesystem design itself. By examiningthese criteria, we can establish thehydraulic, mechanical and materialsdesign requirements that the selectedpump must provide. Once theserequirements are fixed, it is possibleto determine the type of pump thatcould be considered in the applica-tion. This is usually the time todecide whether to use a positive dis-placement or a kinetic (usually cen-trifugal) machine.

For example, pumping a lowflow rate of a highly viscous clearliquid at high pressures would indi-cate selection of a positive displace-ment pump. Conversely, higher flowrates of a low viscosity clear liquid atmodest pressure would suggest acentrifugal impeller pump design.There are many other considerations– pulsed versus continuous flow, sin-gle or multi-stage centrifugals, slurryor solids bearing versus clear liquids,and as previously mentioned, viscos-ity – all of which will influence thetype of pump selected.

Complicated process systemdesigns may require the pump tooperate over variable or several spe-cific flow and pressure conditions.These variations may definitelyaffect the type of pump. Forinstance, an injection pump supply-ing a blending agent at a constantflow rate to a pipe at a location ofvarying pressure would be an appli-cation for a positive displacementdesign.

Commercial pumps are avail-able for most low viscosity clear liq-uid applications. The figure from mypump training course is a headcapacity chart illustrating the rangeof commercially available singlestage centrifugal pumps of the endsuction and axially split (double suc-tion) designs. Note that the versatilevertical turbine design may be usedin the same range of flow and heads

in single or multi-stage design.Portions of the higher pressure area,including those marked for a positivedisplacement selection, may be suit-able for special designs, such as highspeed centrifugal and regenerativeturbine pumps.

Increasing viscosity ofNewtonian type liquids requires spe-cial consideration.

Non-Newtonian liquids are aspecial case not included in this dis-cussion.

Centrifugal pumps may beused for viscous liquids. I haveused 5000 SSU as a maximumallowable viscosity for centrifugalpump application. Increased viscos-ity has a measurable effect oncapacity-head performance andseverely affects hydraulic efficien-cy.

Positive displacement pumps aremore tolerant of viscous liquid appli-cation and are normally used for thistype of service. The hydraulic effi-

ciency of a positive displacementpump, depending on the type, is littleaffected by increasing viscosity atmodest pressures.

It becomes more apparent as thepump selection process progressesthat the properties of the pumped liq-uid yield important selection criteria.■

Until next time ...

Selecting a Pump-What Type Should It Be?

BY J. ROBERT KREBS, P.E., CONTRIBUTING EDITOR

TPosit

ive

Displac

emen

tCentrifugal Commecial Sizes

VerticalTurbine

End Suction

AxiallySplitCase

Head

in F

eet

1,000

100

10

110 100 1,000 10,000

Flow in GPM

FIGURE 1. PUMP TYPES FOR LOW VISCOSITY LIQUIDS



HANDBOOK??

Most pump problems occur onthe suction side, rangingfrom difficulties in suctionpumping to sizing and lay-

out. Suction pressure can also beproblematical—and that is the areawe look at this month.

There is no reason to have thatenigmatic acronym NPSH (NetPositive Suction Head) a part of thesuction pressure puzzle, when it is avaluable clue to it!

There are two values of NPSH.

NPSHA (A for Available) is the pres-sure you, as the designer, provide inyour system. It must be more thanthe NPSHR (R for Required at theanticipated flow rate). The pumpmanufacturer provides the NPSHRon curves for the selected pump.

The value of NPSHR for a givenpump increases with flow rate in theusable centrifugal pump operatingrange. NPSHR decreases somewhatas the impeller eye diameter isincreased (larger suction). NPSHRincreases with operating speed (rpm).

Figure 1 sketches a simple appli-cation with the formula for calcula-tion of NPSHA. A centrifugal orrotary pump is applied here, sincereciprocating positive displacementpumps require a somewhat differentapproach.

Now that you have calculated

your NPSHA, how much does thepump require? You could start leaf-ing through manufacturers’ catalogssearching for a pump that meetsyour need—or you could try anotherapproach.

Many years ago it was discov-ered that the parameter suction spe-cific speed was related to theimpeller design NPSHR. Suction spe-cific speed has the same variablearrangement as specific speed withNPSH replacing system head. So,how can suction specific speed beused to help you in designing yoursystem?

The Hydraulic Institute Stan-dards (14th ed., Pg. 107) features achart of NPSHA versus capacity forvarious operating speeds at a suctionspecific speed of 8500 (Englishunits). The value of 8500 is consid-ered to be an attainable design num-ber for commercial sizes of radialand mixed flow impeller designs. Ishould note that many manufactur-ers produce higher suction specificspeed design impellers, usually clas-sified as “low NPSH impellers.”

This discussion is limited topumps available as a commercialproduct for general application. TheHI chart noted above is for singlesuction impellers. If a double suction

pump is to be selected, the capacitymay be doubled. For example, it a2000 gpm single suction pump at1800 rpm may require 20 ft NPSHR,the same 20 ft requirement in a dou-ble suction pump would permit aflow of 4000 gpm.

I have used this chart for manyyears, both in my consulting practiceand training courses. I have foundthat for a given pump selection it is areliable guide to how much NPSHAthe design will need — and I do nothave to search through manufactur-ers’ curves.

Consider an example with waterat 2000 gpm, 150 ft total head and 24ft NPSHA. From the selection chartin last month’s column, an end suc-tion single stage, single or double suc-tion, or vertical turbine single ormulti-stage pump should be avail-able. At a speed of 3600 rpm (syn-chronous speed), the HI chart showsa requirement of just over 50 ft for allsingle suction selections. A doublesuction pump, (using half the capaci-ty 1000 gpm), would require over 30ft. Reducing the speed to 1800 rpmwould reduce the NPSHR to about 20ft, an acceptable alternative.

Just trying to make pump selec-tion a little easier. ■

Until next time...

Selecting a Pump — Suction PressureBy J. Robert Krebs, P.E., Contributing Editor

There is noreason to havethat enigmaticacronym NPSH

a part of the suction

pressure puzzle.

hp

hvp

hp

hvp

-h

hf

+h

FLOODEDSUCTION

SUCTIONLIFT

hvp = liquid vapor pressure (ft)hf = friction loss at flow rate

NPSHA = hp - hvp ± h - hf (units ft)

CL CL

FIGURE 1. NET POSITIVE SUCTION HEAD AVAILABLE



HANDBOOK??

urrent market forces oftendictate significant downsiz-ing and cost reductions.With the resulting scarcity

of resources for expansion andreliability upgrades, industrialprocessors continually seek meansto maximize capacities with exist-ing equipment. Under these condi-tions optimum performance ofrotating equipment is a must!

Deterioration in pump perfor-mance and the subsequent eco-nomic consequences of productionloss can be addressed with little orno capital expenditures by apply-ing knowledge of pump and sys-tem design principles. A recentexperience with the performancedeterioration of a 1 1/2 x 2 ESN 2stage, top suction top dischargedesign Wilson Snyder pump(Figure 1) provides a good exam-ple of the value of this know-how.Operating at 115 gpm and 890 ftof differential head (TDH), thepump was not duplicating theOEM’s test curve performance(Figure 2). The pump was intend-ed to operate at 115 gpm and toproduce 925 ft of TDH.

After a few simple calcula-tions, the economic incentives topursue a solution to this problembecame clear. The lower head ofthe trouble condition was costing60 gpm of flow. Since the productis sold at $0.02/lb, a calculationcan be made based on the fluidproperties:

COST of head loss:(.68) (62.4 lb/ft3) (1ft3/7.48 gal)($.02/lb) (60 gal/min) (1440 min/day)= $ 9,802.47/day

where:.68 = specific gravity of the fluid62.4 lb/ft3 = density of water

The potential for nearly$10,000 in additional revenue perday justified our efforts to improve

pump performance using expensebudget funds.

Through a brainstorming sessionand careful analysis of the pumpoperating symptoms and system con-ditions, the items listed in Table 1were identified as possible causes forthe performance degradation. Thislist was used as a guide for theinspection and testing of the pump.

When field testing a pump, con-sider the following key factors: • All gauges should be calibrated,

even new gauges. New gaugescan be as much as 25% off cali-bration.

• Gauge readings must be cor-rected to the centerline of thepump suction.

• When using computer-collect-ed flow readings, note theactual temperature and pres-sure of the product and cor-rect the readings fromstandard temperature andpressure conditions to actualconditions. In hot bottomsservices, the flow correctionsare especially significant.

• Measure the running speed ofthe pump for each set of read-ings. These must be corrected

Troubleshooting Pump Performance DegradationBY LUIS F. RIZO

C

OEM CurveProblem Report

1000

950

900

850

800

750

7000 25 50 75 100 125 150 175 200

Capacity (gpm)

TDH

(ft.)

FIGURE 2. REPORTED FIELD PROBLEM

FIGURE 1. TWO-STAGE TOP SUCTION TOP DISCHARGE PUMP


to the OEM’s curve nominalspeed as printed on the per-formance curve.Manufactur-er’s curves are corrected to anominal speed through therange of the performancecurve.

• Verify the impeller diameterto correct for differencesusing the fan laws.

Figure 3 shows the results ofthe field test. These data revealeda significant reduction in thepump’s head as the performanceapproached 30 gpm. This drasticpressure drop with flow increaseis generally indicative of anobstruction to the flow of liquidsomewhere downstream from theimpeller(s). An obstruction’s resis-

tance increases as the square rootof the flow rate. Consequently, atshut-off, when the flow rate iszero, the TDH matches that of theOEM test curve.

Table 2 describes conditionsidentified as possible causes forthe pump performance exhibitedby the field test along with corre-sponding evidence obtained bymechanical inspection or testing.From these findings, the signifi-cant and rapid changes in pipingdiameters near the discharge noz-zle appeared to be the major con-tributor to the sharp drop inpressure. A modification of thepiping is the recommended solu-tion.

Good engineering flow mea-surement practice (Ref. 2) recom-mends that a straight run of 5 to10 pipe diameters be designed atthe discharge of the pump toallow for the process of recovery.Piping changes and restrictionsinterfere with the process of con-verting velocity head into pressurehead, and this effect continues tooccur at up to 5 to 10 piping diam-eters from the discharge nozzle.

In this case a 2”x4” reducer,4” check valve and 4” block valvewere removed and then re-installed downstream of a 10 pipediameter straight piping run. Afterthese modifications were made,the pump was re-tested.

Likely Causes Comments

1. Internal recirculation due to improper The pump is operating and controlled operations at or near BEP. Internal recirculation is

unlikely.

2. Ring clearance not to specifications Possible cause. Field testing reveals that the head degrades with flow, and delivers OEM performance at shut-in conditions.

3. Obstruction in piping system Poor piping design can cause severeperformance problems. This pump hasa significant reduction in the suction and discharge piping.

4. Damaged impeller Not likely. The pump flow is off the OEMcurve, but delivers test curve head at shut-in conditions.

5. Undersized throat or reduced Possible cause. If build up/foreigninterstage passage material is lodged between stages,

significant performance reductionoccurs.

6. System design created obstruction The discharge piping was reduced from4 to 1-1/2 in. at less than 5 pipe diame-ters from the discharge nozzle.

7. Use of non-OEM parts Possible cause, but not likely. Recordsindicate the existing impeller is OEM.

8. Running speed does not match test The pump is directly coupled and thespeed running speed is close.

9. Errors in factors used to correct flows Possible cause. When adjusting flows toto STP STP the operating temperature effect on

the material density must be taken intoaccount.

TABLE 1

OEM CurveProblem ReportField Test

1000

950

900

850

800

750

7000 25 50 75 100 125 150 175 200

Capacity (gpm)

TDH

(ft.)

FIGURE 3. FIELD TEST VS. OEM CURVE


The results, shown in Figure4, demonstrate a significantimprovement over the previousperformance. Throughput wasincreased by 30%, and due to sys-tem characteristics the additionalflow was directly translatable toadditional product to sell. This per-formance improvement wasattained using a very smallamount of business funds relativeto the cost of replacing the pump.However, every system is not thesame, and other systems mayrequire additional investments oftime and resources to producesimilar results. Foremost, youmust be familiar with the system,its operating mode and how it iscontrolled. ■

REFERENCES

1. Yididiah, S. Y. CentrifugalPump Problems, Causes andCures. Petroleum PublishingCo., 1980.

3. Cheremisinoff, N. P. FluidFlow, Pumps, Pipes andChannels. Ann Arbor Science,1982

Luis F. Rizo is ReliabilityEngineering Manager for G.E.Silicones in Waterford, NY. He isalso a member of Pumps andSystems User Advisory Team.

Description of Condition Discussion of Applicability

1. Undersized throat of the volute or This condition was not present whenreduced interstage passage the casing was inspected. No build-up

or blockage was found.

2. Ring clearance not to specifications The rings were inspected and theclearance measured/restored to OEMrecommendations prior to re-installing.

3. Obstruction in piping system No physical evidence of obstruction wasfound. However, as mentioned previous-ly, the discharge piping was reduced from 4 to 1-1/2 in. at less than 5 pipe diameters from the discharge nozzle.

4. Use of non-OEM parts Inspection revealed all serialized OEMparts.

5. Errors in factors used to correct flows All corrections made to the data reflect to STP actual conditions.

TABLE 2

OEM CurveProblem ReportField TestTDH After Mods.

1000

950

900

850

800

750

7000 25 50 75 100 125 150 175 200

Capacity (gpm)

TDH

(ft.)

FIGURE 4. PERFORMANCE AFTER MODIFICATIONS



HANDBOOK??

AT THE JOB SITE, 8 A.M. Unit 4 is off the line for a reg-

ular maintenance check. Bill andMike, the maintenance engineers,take me to the installation andthen to the shop where a com-pletely assembled unit is ready formounting. We discuss the mainte-nance records—not one 600 hpvertical turbine condensate pumphas operated for a full year with-out a failure. Resonance is ourfirst suspicion.

The structures used forpumps and their mountings havenatural frequencies that causevibration and noise amplificationwhen excited by a resonant fre-quency. My experience consult-ing on a case at Pacific Power &Light in Glenrock, WY helps toillustrate the excessive mainte-nance problems created by self-induced pump resonance.

NOT SO MUSICAL

A tuning fork will always res-onate at the same tone, and sameovertones, when struck. This istrue of all rigid structures.Although natural frequency reso-nance is employed in musicalinstruments and in our own vocalcords, it must be avoided in rotatingmachinery applications. This con-trast in attitudes toward resonancefrequencies is evident even in thelanguage we use to describe them.Frequencies referred to as themajor tone in music are called thefirst critical for pumps. Likewise,the first overtone in music is thesecond critical in a pump structure.

Table 1 lists some sources ofvibration energy in pumps.Residual unbalance and impellervane passing (impeller vane pulsewhen passing a cutwater guide ordiffuser) are the two primary caus-es of exciting forces in centrifugalpumps. Spring mounting of recip-rocating equipment is often select-ed so that the first critical naturalfrequency is only 10% of the oper-ating rpm. However exciting

forces in centrifugal pumps are rela-tively low, and the natural frequencymay approach 75% of an excitingforce. As a rule, avoid the first criti-cal by +/– 25% and the nth criticalby +/–[(25)1/n]%. The followingexpression can also be applied:

Fn (.75)1/n ≤Rn≥ Fn (1.25)1/n

where: Rn = frequency range to avoidfor the nth critical

Fn = n critical frequency

MEASURING NATURAL FREQUENCIES

Back at Unit 4, our first step wasto measure the natural frequencies ofthe pump installation. To facilitatethe test, we installed the pump withthe discharge head set on a woodblocking about two feet off the can.This set-up gave us access to the col-umn with a rubber hammer. Thelowest critical frequency we coulddetermine was the second. Table 2lists the critical frequencies as deter-

mined by our method. A log-logplot of the points confirmed theaccuracy of the measurements.

The third critical, at 1800cycles per minute (cpm), con-firmed our suspicions. This wasthe interfering frequency whichhad caused the 12” column sus-pension failures. These failureshad occurred despite the plant’sstrict observation of alignmentprecautions and the conversion ofthe intermediated shafts to a sin-gle larger shaft.

Resonant frequencies can bedisplaced by altering the stiffnessof the suspension assembly. ForUnit 4, I suggested a three-armedspider between the can and theflange at the bottom of the top col-umn, an arrangement that wouldyield an approximate 23%increase. When I made my rec-ommendation to the people atPacific Power & Light, I received asmall dose of good natured ”static”for proposing a solution thatseemed so insultingly simple.Clearly, these physical principlesare not new to most pump opera-tors, but often their applicationisn’t as obvious.

CONTROL OF CRITICAL FREQUENCIES

Control of static structural,resonant critical frequencies, suchas those experienced with Unit 4,is a field responsibility. Becausethe manufacturer is not generallyinvolved in decisions pertaining tomounting rigidity and alignmentacross couplings, the field design-er must provide the final defenseagainst unnecessary amplificationof vibrations.

On the other hand, control ofdynamic (rotating) structural reso-nant critical frequencies is a manu-facturing responsibility. Becausethree-phase motor drivers areavailable in limited choices of rpm,trial and error over time eliminatemost rotating velocities that mightcause trouble. However, variable

Sources of vibration energy originating the pump or driver

1. Residual unbalance2. Vane pass3. Oil whip4. Misalignment resulting in radial/axial

shock5. Motor windings6. Worn frictionless bearings

TABLE 1

Measured critical frequencies-Unit 4

Critical Frequency Vibration# Frequency (cpm)2 6003 1,8004 3,6005 6,8006 11,5007 18,0008 30,000

TABLE 2

Vibration Amplification


frequency drivers are now becom-ing quite common. These drivescan create unforeseen and exces-sive vibration levels at ”in-between” frequencies. Controladjustments may permit the block-ing of a troublesome frequency,but the required gap often pro-duces undesirable effects on pumpcharacteristics.

The use of inexpensive vari-able speed drives can create costlyproblems in the field, so a thor-ough evaluation of potential vibra-tion sources is recommendedbefore installation. Low rpm,heavy overhung trash pumpimpellers and a limited number ofimpeller vanes, coupled with asingle diffuser/cutwater can pro-duce an expensive coincidence ofexcitation and impeller assembly/cantilever critical frequency ampli-fication. Impeller assembly criti-cal frequencies can be controlledonly at the pump design level.However, I’ve discovered that themathematical computations andempirical field checks on criticalfrequencies are not sufficientlyclose in either foreign or domesticpumps. Whether the volute iswater filled or dry, will make littledifference in measurement of animpeller assembly critical. A vibra-tion analyst can easily make thedetermination either before orafter shipment.

BREAKING CADENCE

The pumps on Unit 4 hadundergone seven years of highmaintenance due to self-inducedresonance previous to our diagno-sis. But, resonant frequencies arenot only damaging for pumps. Formore than 2000 years groups ofmarching men have broken stepwhen crossing a bridge to avoidthe damage possible if the cadenceof their marching might happen tomatch the critical frequency of thebridge span.

ABOUT THE AUTHOR:

Ken Hawkins is the owner ofVibration Control, an independentconsulting firm in Overland Park,KS. He has worked with pumpssince 1950. ■

Ed’s Rules of ThumbEditor’s Note: From time to time, the Shoptech section of Pumps

and Systems will feature Pump Rules of Thumb from Ed Nelson, a notedpump consultant and member of our Editorial Advisory Board. TheseRules of Thumb are derived from a blend of engineering principles andexperience. They are intended to provide a rapid assessment of conditionswhile troubleshooting a pumping system problem, and may be used as asupplement to general or detailed instructions furnished by the manufac-turer. Each machine is different in design or construction but the funda-mental principles apply to all of them.

About the Author: William E. (Ed) Nelson is a turbomachinery consul-tant based in Dickinson, TX. He is the author of more than 50 technicalpapers and a contributor to several handbooks on pump operation andmaintenance. Previously he spent 36 years with Amoco Oil in various engi-neering, materials management and maintenance positions. Mr. Nelson isa registered professional engineer in Texas.

RULES OF THUMB

Pump Operating TemperatureDischarge vs.Nozzle Warming Stream Flow Rate (gpm)Size 200°F 450°F 700°F4”-6” 4 5 68”-10” 5 6 712”-14” 6 8 1016”-18” 7 10 1320” 8 11 14

RULES OF THUMB1. Employ suction piping one or two pipe sizes larger than the

pump nozzle. Suction lines should never be smaller than the pump suction nozzle.

2. To prevent cavitation in the pump, suction line velocitiesshould not exceed 10 ft./sec. Consider 5-6 ft./sec. asa maximum for new systems.

3. Consider the pressure drop across permanent suctionstrainers.

4. Install valve stems and tee branches perpendicular to,not parallel to, the shaft.

5. Employ an absolute minimum of five pipe diameters of straight run before the suction flange. Seven pipe diameters is the preferred minimum.

HOT SERVICE PUMPS—WARMING STREAM FLOWS

A single stage, double suction pump should be warmed up from “cold”to “hot stand-by” by passing hot liquid from discharge to suction for atleast two hours. The warming stream flow rates shown below are aminimum and will require adjustment for a specific installation.

FACTORS IN SUCTION PIPING LAYOUT

Suction piping can be the cause of major pump damage—especiallywith double suction impellers. To reduce these problems with thesuction piping apply the following practices:

A ll users want the highest effi-ciency from their pumps. Thisis why having an accuratedefinition of efficiency prior

to selection is so important.An operating pump transfers flu-

id adding energy (power) to the fluidas it passes through the pump. Thepower developed by the pump(HHP) compared to the power to thedriver (PI) gives the overall efficiencyof the pump and driver.

Figure 1 illustrates a pump anddriver coupled to a system. The dri-ver takes in power. It can be an elec-tric motor, a liquid or gaseous fueledengine, a hydraulic or air powereddriver. The driver has an efficiencythat varies with its power output.The power to the driver (PI) dividedby the power from the driver (PO) isthe driver efficiency (ED).

The coupling between the pumpand driver and any other additionalmechanical components, such as agear or belt drive, also have losses andabsorb some of the driver output pow-er before it reaches the pump. Gearand belt drives and some couplingdesigns have a measurable powerrequirement. The component manu-facturer can provide an estimate ofthe loss. To simplify this discussion,we will ignore losses between the dri-ver and the pump.

The power to the pump providesthe fluid flow and pressure. Theenergy terms of flow and pressurecombine to equal the power out.[Power from the pump divided bydriver power to the pump is thepump efficiency.]

The calculations for efficiencyare indicated in Figure 1. Pump effi-ciency stated by the manufacturerrefers to water at the pumped liquidat STP (standard temperature andpressure) conditions. An increasing

Selecting a Pump? Define Efficiency First!

By J. Robert Krebs, P.E., Contributing Editor

PUMPDRIVERCOUPLING

EFFICIENCY(ED)

EFFICIENCY(EP)

POWER IN (PI)

POWER OUT(PO)

POWER INTO PUMP

(PIP)

FLOW OUTQ2 P2

Q1 = Q2 = Q

FLOW INQ1 P1

ED = PO/PI HP = 0.746KW HHP = Q (P2 - P1) / 1714OVERALL EFFICIENCY = HHP/PI Q (GPM) P (PSI)OVERALL EFFICIENCY = ED X EP EP = HHP/PIP

Figure 1. Power calculations — pump and driver


HANDBOOK?


SPECIFIC SPEED Ns =RPM √GPM

H3/4

Figure 2. Pump efficiency versus specific speed and pump size

RADIAL-VANE FRANCIS-VANE MIXED FLOW AXIAL FLOW

500 1000 2000 3000 4000 10,000 15,000

100

90

80

70

60

50

40

Effic

ienc

y, p

erce

nt

100 gpm

200 gpm500 gpm

1000 gpm

3000 gpm

10,000 gpm

Over 10,000 gpm

CENTRIFUGAL AND AXIAL FLOW PUMPS

viscosity of pumped liquid reducesthe efficiency of all pump types—cen-trifugal more so than rotary types.Centrifugal pumps are used forpumping viscous liquids to about3000 SSU (Seconds Saybolt Univer-sal)—about 600 cSt (centistokes) on aroutine basis. Methods are availablefor approximating the degradationeffect of viscosity on head, capacity,and efficiency (the Hydraulic Instituteis a source for this information).Rotary pumps (gear, screw, circum-ferential piston, and similar designs)when pumping a viscous liquid, givenproper suction and application condi-tions, also have reduced efficiency.Since a more viscous liquid reducesslip, the flow rate and head developedin rotary pumps may be minimallyaffected. Manufacturers should beconsulted for specific applications.

Figure 1 also suggests that thedriver—and pump—efficiency mustboth be considered to determineoverall efficiency. In my experience,too frequently the driver efficiency isoverlooked.

The overall efficiency is the prod-uct of pump and driver efficiency. If,at a certain pumping condition thepump is 80% efficient and the driveris 70% efficient, the overall efficiencyis 56%. With the common electricmotor driver, the overall efficiency iscalled the “wire to water efficiency”(WWE). The WWE is defined as thehydraulic horsepower (HHP) from thepump divided by the measured elec-tric power to the motor—kilowatts(KW) expressed as horsepower (HP).

So what is a good pump efficien-cy? For centrifugal pumps, I have longused the chart in Figure 2 (originallypublished by Worthington), whichdetermines efficiency from pump size(flow) and impeller shape as predictedby specific speed. This chart teachesthat a single stage (or the first stage)centrifugal pump becomes more effi-cient as flow increases for radial vanedesign impellers.

Now let’s take a look at the chart(remember—low viscosity liquidsonly, such as water). For a pump rat-ing of 200 gpm at 100 ft total head,

operating at 3500 rpm, the specificspeed calculates to 1565 and predict-ed efficiency is 70%. [Note theimpeller design is radial vane.] If theconditions were 2000 gpm at 100 fttotal head and 1750 rpm, the specificspeed is almost 2500, and the estimat-ed efficiency would be 80% or more.The impeller design is in the Francisvane area of radial flow design.

Rotary type pumps are quite effi-cient, 70–80%, when pumping lowviscosity liquids. They adapt well topumping highly viscous liquids.Manufacturers should be consultedfor specific power requirements.

Being able to estimate efficiencyquickly is useful in preliminary sys-tem design. If the liquid propertiesand flow and pressure are known,the driver size and pump size canthen be approximated along withbuilding sizes and power systemrequirements. ■

Till next time...


Selecting a Pump: Will it Operate Where You Want it to?


You need one or two pumps foran application. The systemanalysis is complete andyou’ve arrived at a flow rate

and pressure (head). Let’s assumethe pumps are constant speed motordriven centrifugals and you, thedesigner, have used a conservativeapproach to your calculations.

The key point—you have decid-ed the Q (gpm) and H (head) know-ing that the pump only operates atthe intersection of its performancecurve with the actual system resis-tance curve.

It is the rare system designerwhose calculations are so precisethat the operating pump actually per-forms at the predicted conditions.

Consider a single source/singledestination open transfer system to apressurized vessel. Both the sourceand destination pressures can differfrom design. The centrifugal pumpmotor driver speed varies slightlywith increasing horsepowerdelivered (centrifugal pumptest curves are usuallymade for a constant speed).The pipe friction lossesvary with type, age andcondition of the pipe. Thepumping of viscous liquids,slurries, paper stock andsimilar combinations is notconsidered in these com-ments.

Figure 1 illustrates aperformance curve from amanufacturer’s catalog.Super-imposed on it are aschematic of an open trans-fer system and calculatedsystem head curves A, Band C. Curve A intersectsthe maximum diameterimpeller performance at thedesign flow rate and head.

Curve A was constructed with maxi-mum static head and a pipe resistancecoefficient equivalent to aged (14–17years) pipe. As the pipe ages in use,resistance to flow normally increases.

Curve B represents new pipe atthe minimum expected static head.Curve C is for the same static head asB, but with the pipe aged as in A.

Assuming that the system designcalculations for head are accurate,and the pump speed is as shown onthe curve, the pump will operatewith new pipe at the curve B inter-section corrected for the actual statichead. As the pipe ages, the intersec-tion point will be between the curvesA and C. In other words—there are anumber of system curves with pro-gressively higher heads and com-mensurately lower flows.

How do the pumps and systemsinteract? From the performancecurve, as flow increases, so do brakehorsepower and required NPSH (Net

Positive Suction Head).Another limitation is available

suction head (NPSHA). If the head isoverstated, the pump will run at B,rather than A in Figure 1. Is thereenough NPSHA? What horsepoweris required? Low to medium specificspeed design centrifugal pumps havean increasing horsepower require-ment with flow rate. If the systemresistance is lower than calculated,will the pump require a higher suc-tion pressure or a larger motor?

These can be big problems witheven small pumps. How can this becontrolled?

To start, the system designer mustbe very sure that all pump and pipingelevations are correct. Next, thedesigner in his calculation of systemresistance should use both new pipeand aged pipe resistance factors. Theaged pipe factor must reflect the antic-ipated actual life of the system. Pumpselection and motor (driver) sizing

FIGURE 1

U.S. GALLONS PER MINUTE

0 400 800 1200 1600 2000 2400

200

160

120

80

40

0

FEET

30

20

10

0

12

8

4

0

FEET

MET

ERS

NPSH

-R

CUBIC METERS PER HOUR

300 400 500

NPSH-R

B

C

L

L

11″

10″

9″

20 HP25 HP

30 HP40 HP

50 HP

60 HP

60%

71%74

%76

%

78%

79%

77%76

%

74%

A

78%

77%

64%68

%

79%

12″

MET

ERS

60

50

40

30

20

10

TOTA

L HE

AD

L= Limit Line

Model6312-3D

1750 R.P.M.40200

Impeller No.

Y-4677Number of Vanes

2Max. Sphere

3″Discharge Size

6″Suction Size

6″ or 8″Inlet Area

28.27 sq. in.


HANDBOOK??


must reflect these concerns consistantwith the anticipated life of the system.

There are many other operatingrange problems that cannot be dis-cussed in just one column. To men-tion a few:

1. What do minimum/maximum

flow rate lines on a performancecurve mean?

2. How does one determineNPSHR (with flow)?

3. When do motors overload?4. How accurate is the manufac-

turer’s estimate of functions

(NPSHR–efficiency)?Each of these questions deserves

a lengthy answer. Maybe we canaddress them later. ■

Until next time...


Apump’s purpose is to transfermechanical energy from amotor into hydraulic energywithin a fluid. This energy is

then dissipated as the fluid movesthrough a system. The pump’s headcapacity (H-Q) curve defines howmuch energy is available at a givenflow rate and shaft speed. A samplehead capacity curve at 1800 rpm,supplied by the vendor, is shown inFigure 1. For convenience, the ener-gy level or head is expressed in unitsof feet.

A simple system is shown in Fig-ure 2. Only friction and velocity head

losses are present in this system.Velocity head losses occur when thearea in the flow loop changes sudden-ly. In this simple arrangement thesystem head curve is zero when thereis no flow.

Figure 3 shows a similar systemwith a net change in elevation. Inthis case the pump must provide anadditional amount of energy equal tothe change in elevation before flowcan begin. This difference in eleva-tion, expressed in feet, is identifiedas hs and is normally referred to asstatic pressure.

The “normal” system head curve

shown in Figure 1 represents theenergy required to overcome staticpressure, friction and velocity headlosses present in a system. The sys-tem head curve establishes thepump’s operating condition. Thepoint where the system head curveand the pump’s head capacity curveintersect defines the point where thepump will operate at a given speed.This point is identified as A1 in Fig-ure 1.

For a pump to operate in a givensystem, it must be capable of deliver-ing a head that is greater than thesystem static pressure (hs). Note in

Performance Curve vs. System Curve

By Phil Mayleben

Normal System Head CurveThrottled System Head Curve

Pump H-Q @ 1,800 RPM



Pump Best Efficiency Point

C1

A1

A2

A3

B1

B2

B3

HEAD

IN F

EET

FLOW IN USgpm0 100 200 300 400 500 600 700 800 900 1000 1100 1200 1300 1400 1500

Friction (h1 - hs)

Static Pressure (hs)

100

90

80

70

60

50

40

30

20

10

0

FIGURE 1. Sample head capacity curve


HANDBOOK??


Figure 1 that if the pump’s shaftspeed is reduced from 1800 rpm to1600 rpm, the head and flow capabil-ity decrease. If the speed of thepump is reduced much below 1200rpm, it will no longer be able topump in this system because its headis approaching 30′, which is equal tothe static pressure (hs).

The pump Best Efficiency Point(BEP) occurs when the ratio ofhydraulic power to input power is ata maximum. The BEP for the pumpin Figure 1 occurs at 1000 gpm and70′ at 1800 rpm (point B1, and thendrops to about 667 gpm and 31′ at1200 rpm (point B3).

Note that at 1800 rpm the oper-ating point (A1) occurs at a flow thatis greater than the pump’s BEP (B1)).As the speed is reduced to 1600 rpm,the BEP (B2) and operating point (A2)are almost identical. By the time thespeed is reduced to 1200 rpm, theoperating point (A3) is significantly tothe left of the pump’s BEP (B3).

The system head curve shouldbe carefully calculated by the sys-tem designer. If it is in error, thepump may not operate at the intend-ed duty point. When possible,pumps should be selected so that theBEP is close to the system headcurve. In Figure 1 this occurs atabout 1600 rpm. Operating a large,high power pump too far away fromits BEP can cause pump damage,excessive wear and high vibrationlevels. The risk of damage is reducedas speed is decreased. The locationof the pump’s BEP, and recommen-dations on how far away from BEPone can operate a pump safely,should be supplied by the pumpvendor.

If a pump operating point basedon Q1, H1, N1 and HP1 is known,then a new operating point at a newspeed N2 can be estimated. The fol-lowing relations, known as the affin-ity laws, can be used to step knownpump performance to a new speed.

Flow:Q1/Q2 = N1/N2

Head:H1/H2 = (N1/N2)2

Power:HP1/HP2 = (N1/N2)3

In these equations Q is thepump flow rate, H is the pump head,N is the shaft speed, and HP is thepower required by the pump. Sub-script 1 refers to known values, andsubscript 2 applies to calculated val-ues.

The curves at 1600 and 1200rpm in Figure 1 were calculated fromthe 1800 rpm curve by using theseequations.

It is also possible to calculate asystem head curve similar to Figure1 if a single flow-head point isknown. At point A1 in Figure 1, thetotal system head (h1) is 67′. The sta-tic pressure (hs) is 30’. The frictionalpart of the system head curve (h1 - hs)is 37′. Because frictional and velocityhead losses are proportional to the

square of the flow rate, an equationfor calculating a new point (h2) onthe system head curve can be writ-ten as:

h2 = (q2/q1)2 x (h1 – hs) + hs

Note that lower case letters areused for q and h in this equation sothat the system head and flow valuesare not confused with pump parame-ters. The equation assumes that hs isconstant and would not apply to asystem with variable static head. Foran example of variable static head,consider a system in which thepump is used to empty a tank.Because the elevation of the waterfeeding the pump is constantly drop-ping as the tank is drained, thewhole system head curve increases

SUCTION VALVE

FLOW

DISCHARGE VALVE

VELOCITY HEAD LOSS

FIGURE 2. Simple pump system

SUCTION VALVE

FLOW

DISCHARGE VALVE

hs

FIGURE 3. Pump system with elevation change


as the pumping proceeds. The rate ofchange in hs depends on the geome-try of the tank.

The most efficient way todecrease the flow delivered by apump is to reduce the shaft speed.This is easy if some sort of variablespeed control device is available.Unfortunately, many pump installa-tions are driven directly by constantspeed AC induction motors. In thissituation the only way to adjust thecapacity is to increase the systemhead against the pump artificially bythrottling a valve on the dischargeside of the pump. While this address-es the need to reduce flow, it alsowastes energy.

Referring again to Figure 1, it ispossible to force the pump’s operat-ing point from A1 to point C1 by par-tially closing a valve on the dischargeside of the pump. The pump thenproduces a head of about 71.5′ at 910gpm in front of the discharge valve,but the difference in head betweenpoints C1 and A2 is lost across thepartially closed valve. Thus, only theenergy available at point A2 reaches

the system. Although this flow con-trol method works, it is comparableto turning your home heating systemon full blast and then regulating theinside temperature by opening thewindows! It should be reiterated thatmany large high-power pumps can-not be operated at low flows withoutsustaining damage.

The following additional pre-cautions should be observed whenpump speed changes are anticipatedin a system. First, be aware that highvibration levels could develop at cer-tain speeds if you are unfortunateenough to encounter a natural fre-quency. Vibration problems due to anatural frequency are more commonin vertical pump units. If the pumpvendor is made aware of the needfor variable speed pumping capabili-ty when the equipment is pur-chased, he can offer suggestions toavoid any possible natural frequen-cy problems.

If anticipated speed increase isnecessary, one must determine thatthe pump is adequately designed towithstand the increased loads, and

that sufficient horsepower is avail-able from the pump driver. Notice inthe above equations that horsepowerincreases with the third power of thespeed ratio.

An increase in speed alsorequires that an adequate suctionpressure or NPSH (net positive suc-tion head) margin is still available forthe pump. Inadequate NPSH canresult in significant loss of perfor-mance as well as noise, vibration andimpeller damage. Usually, pumpNPSH requirements can be expectedto increase with the square of thespeed in the same manner as head.Again, the pump vendor should beable to assist in these areas. ■

REFERENCES1. Stephen Murphy, “Variable

Speed Pumping,” Pumps and SystemsMagazine, April 1993.

2. A. J. Stepanoff, Centrifugal andAxial Flow Pumps, 1957, John Wileyand Sons, Inc.

Phil Mayleben is employed by ITTA-C Pumps (Cincinnati, OH).


The article on shaft stiffness fac-tor calculations by Dan Besicin the February Pumps and Sys-tems caught my attention. He is

“right on” in considering all of thefactors in the shaft stiffness formulawhen comparing the relative shaftstiffness of two or more similarpumps.

It is understood that the deflec-tion formula for a simple cantileverbeam (shaft) is y=FL^3/3EI or sinceinertia I=πD^4/64 for a round shaftand the modulus for steel is E=28 to30×10E6psi (I use the larger num-ber), then y=FL^3/44.20D^4 (y ininches ×10E–3 or mils). This formuladoes not account for the overhung

shaft weight and the shaft length anddiameter between the bearings,which also contributes to shaftdeflection in overhung impellerdesign (end suction and double suc-tion) pumps.

The shaft stiffness factor is aconvenient short-cut for comparingthe relative stiffness of two pumpshafts provided the materials of con-struction and the other design crite-ria, such as shaft diameters, areidentical or essentially similar, as inANSI pumps.

Use of the shaft stiffness factorhas become popular for comparingtwo ANSI pumps of the same size.Perhaps a review of the use of shaft

stiffness factor will be sparked byrecent advent of several so-calledsuper ANSI designs.

For more than 40 years in thepump business, I have used a simpleformula that includes both bearingspans and their shaft dimensions,along with the Hydraulic Instituterecommended radial thrust factor tocalculate shaft deflection.

As chairman of the HI TechnicalCommittee on radial thrust, I over-saw original publication of this workin the Hydraulic Institute Standards12th Edition (1969). The latest HIStandards (1994) has expanded thiswork (pages 103-105) and is thesource I now use for this discussionon shaft deflection.

Today, most pump manufactur-ers use a modified finite elementanalysis (FEA) approach for the shaftdeflection calculation. The formula Iuse (Figure 1) produces a slightlylower deflection than the FEAmethod.

I calculated the maximum (zeroflow) shaft deflection at the impellercenterline for two manufacturers’4×3×10 ANSI pumps at 3500 rpmusing the formula in Figure 1. Pumpmanufacturer D-1 calculated to 2.6mils and for manufacturer D-2 a val-ue of 1.8 mils. Since the face of thestuffing box is about one-half the dis-tance from the centerline of theimpeller (maximum deflection) tothe inboard bearing (zero deflection),the relative values are about 1.3 and0.9 mils respectively at the face ofthe stuffing box. Manufacturer D-2published data at the face of thestuffing box is 1.1 mils.

For horizontal shaft pumps theweight of the impeller must be addedto the calculated radial thrust load.In this example, the impeller weightof 13.6 lb increases deflection byalmost 20%.

Does giving effect to the shaftlength and diameter between bear-ings make a substantial difference inshaft deflection? I estimated the max-

Calculating Shaft DeflectionBy J. Robert Krebs, P.E., Contributing Editor

R1

R2

D1 D2

Y

P

L1 L2

xStuff box face

RPM

P L1^2 L1 L2

4420 D1^4 D2^4Y = [ ]+ Yx = Y

L1 - x

L1

(approx.)

L & D InchesP PoundsY Inches x 10E-3 or Mils.

FIGURE 1. Calculating shaft deflection


HANDBOOK??


imum radial thrust for the 4×3×10 at71 lb; adding this to the 13.6 lbimpeller weight increases the stiff-ness factor deflection to 1 mil for thestainless steel pump in the Februaryarticle or exactly one-half of the totaldeflection calculated from the for-mula in Figure 1.

I would submit that using adeflection formula that includesshaft dimensions between the bear-ings is justified.

What is not understood some-times by the pump user is the impor-tance of specifying maximum shaftdeflection on a purchase request orrequest for quote (RFQ).

The conventional volute casing

centrifugal pump has the highestradial thrust load; therefore, it hasthe largest shaft deflection at zeroflow and practically zero radialthrust load at the best efficiencypump flow.

Excessive shaft deflection, morethan 2 mils at the face of the stuffingbox (packing or seal gland), will short-en packing or seal life, subject theshaft to cumulative reverse bending(fatigue) stress, that could cause shaftfailure and increase the bearing loadwith a commensurate reduction inbearing life while elevating inboardbearing operating temperature.

So what’s the answer? Control-ling shaft deflection and bearing

loads to appropriate levels is, in myopinion, the most important factor inassuring a quality pump mechanicaldesign. The RFQ to the suppliershould ask for calculated shaftdeflection and bearing life calcula-tions for the pump proposed at theoperating condition (or an operatinghead range).

If you want to put it into moreprecise and detailed specifics, dropme a line. I will send the details sothat you can properly specify shaftdeflection and bearing life to assure aquality pump mechanical design. ■

Until next time ...


Last August this column dis-cussed pumps operating in par-allel. Jim Marean of FluidKinetics in the Buffalo area

asked me to introduce variable speedinto the discussion of parallel opera-tion. I also received several callsrequesting amplification of mini-mum and maximum flow restrictionsfor centrifugal pumps and how theyaffect the pump. These are inter-con-nected questions. This discussionwill concern itself initially with two(or more) identical centrifugalpumps operating from the samesource, with each motor driverequipped with a variable speed fre-quency drive control.

Figure 1 illustrates two identicalpumps, each operating at full speed.Note that when the two pumps oper-ate in parallel, each one producesless flow, but together the two pro-duce more flow than one pump oper-ating by itself. Figure 1 shows theeffect of a high friction loss system“A” on the total flow of two identicalpumps in parallel. Note that as thedesign system friction loss decreases(system B & C), the contribution tototal flow from adding a secondpump increases.

What is illustrated in Figure 1 isapplicable to all centrifugal pumps.When special impeller designs, suchas non-clog and solids handling, areconsidered, operation at reducedflow must be carefully analyzed. Themanufacturer may recommend aminimum flow in these designs as apercentage of capacity at Best Effi-ciency or incorporate a minimumflow line on the performance curve.

Minimum and maximum flowlines, when noted on performancecurves by the manufacturer, restrictthe pump flow to that operating

range to avoid problems. If operatedbelow the minimum recommendedflow, increased noise, vibration andrecirculation, as well as increasedradial thrust with accompanyingincreased bearing loads, will assureincreased bearing operating temper-atures and shorter bearing life. Wecould also mention increased shaftdeflection with shorter seal life orincreased packing leakage and shaftbreakage.

Figure 2 illustrates the maxi-mum and minimum speed of one oftwo identical pumps and the systemhead curve against which they mustoperate at some speed between min-imum (N1) and maximum (N4). Thecontrol system is programmed tostart a second pump, also variablespeed-driven.

There are two control-operatingoptions, lead-lag or load-share. Inlead-lag, the speed of the second orlag pump is speed-adjusted by thecontrol system to handle only addi-tional flow from the source.

In Figure 2, if the lead pump isoperating at speed N3 when the lagpump is activated, the latter willincrease in speed to some speed, sayN2, at which point it will start pump-ing the excess flow that the leadpump cannot handle. The curve Arepresents the performance of thelead pump at speed N3 plus the lagpump at speed N2. The horizontalline notes the flow from the lagpump, also noted as QA, less thanthe minimum flow line. This is oneof the dangers of lead-lag operation.

In this example, in the lead-lag

Pumps in Parallel with Variable Speed Drive


FIGURE 1. Two pumps in parallel


HANDBOOK??


Hea

d F

t.

A

B

C

Pump 1 + 2

Pump 1

Contribution of Second Pump to Total Flow

Flow Q

mode, the lag pump must have aminimum speed in excess of N2 toassure minimum flow in excess ofthe minimum flow line.

Of course, the same restrictionapplies to the lead pump. That is, thecontrol system must recognize theminimum variable speed setting foreach pump to assure operation inexcess of the minimum flow value ateach speed.

Some lead-lag systems operatewith the lead pump attaining maxi-mum speed before a lag pump isstarted. In those cases, most often thelead pump continues to operate atmaximum speed while the lag pumpvaries in speed to pump the excessflow. The same comments apply –

the lag pump speed must assure flowin excess of the minimum flow line atall speeds. If the pump performancecurve being considered does notshow a minimum flow line, questionthe manufacturer. For non-clogdesigns, both radial and mixed-flowimpeller, there is a minimum flow.

Load-share is my preferredmethod of operation.

Referring again to Figure 2, atsome selected speed (N2), the lagpump joins the lead pump in opera-tion. Both pumps will operate at thesame speed to maintain the sourcesignal control flow. The curve B inFigure 2 represents the performanceof both pumps at speed N2. The flowfrom each pump is QB – outside the

minimum flow line. If the source sig-nal calls for more flow, both pumpsrespond with increased speed to sup-ply the flow. If less flow is required,both pumps slow down.

At some lower speed, when bothpumps are approaching the mini-mum flow, one pump will shut offand the remaining unit will adjust itsspeed to the flow signal.

At station design, pump sizeselection and load-share may be usedto operate the station pumps at theiroptimum efficiency, minimizingpower costs.

Where considerable variation inflow occurs at a pumping station,more than one size pump may beinstalled. It is important that the par-allel operation of unequally sizedpumps, whether variable or constantspeed, be checked out to assure theirproper operation in parallel.

What about maximum flow lineson performance curves? Flow sub-stantially beyond the Best EfficiencyPoint capacity, commonly called a runout condition, introduces increasedNPSHR, may require additionalhorsepower and frequently producessubstantial noise and vibration.

Centrifugal pumps should not beoperated at severely restricted or runout conditions. Depending on thedesign, specific speed and type ofpump, an operating range of 25-110%of BEP capacity is a rule of thumb forsmaller, say through 8″ dischargesize, radial flow impeller designpumps. As pump size increases, theallowable operating flow range fromBEP becomes more restricted. Formixed-flow design impellers, theoperating flow range may be from 45-105% of BEP capacity for 14″-18″ andlarger discharge sizes. ■

Until next time ...

FIGURE 2. Variable speed pumps in parallel—lead-lag and load-share operation

Min. Flow Line

Head

Ft.

SystemResistance

Max. Flow Line

QA QB N1N2

N3

N4A

B

A = Lead-LagB = Load-Share

Flow Q


Sizing Pumps for Complex Systems: Part IBy J. Robert Krebs, P.E., Contributing Editor

For the pump system designer,the rare case is an open transfer,single-source single-destinationsystem. If only real life were

always that simple! In practice, a mul-ti-source single-destination systemsuch as the one shown in Figure 1 iscommon. The system is descriptive ofa blending or injection system, wheresource 2 is adding a predeterminedamount of material to the source 1flow. Some examples would be pHbalancing, polymer feed addition orthe introduction of blending agents.Pumps in these services can be cen-trifugal or positive displacement or acombination, as determined by flowand pressure, and obviously, eithersource flow can vary.

For the experienced systemdesigner, the problem in Figure 1 isfamiliar, but I will use it in explaininga manual approach that can beapplied to solving difficult problems,such as those that follow in laterissues of this column. The manualsolution is a desired method in thisinstance. Expert pipe network pro-grams fast become a more expedientapproach as the system complexityincreases, but I have always liked thecozy feeling of knowing how to solvethe problem manually. Computer pro-grams are only as good as the datasupplied and the ability of the user toanalyze the results. Perhaps learninghow to approach the solution of acomplex system problem will increasethe less experienced designer’s confi-

dence in using an expert program.In Figure 1 the problem is sizing

pumps 1 and 2. The variables areflow rates (Q1 and Q2), the dischargepressures (P1 and P2), the destina-tion pressure (P3), and the liquid ele-vation differences (and pressure for aclosed system).

Z1, Z2 and Z3 elevations are froma datum reference elevation (normal-ly the shaft centerline of one of thepumps). Pressures P1 and P2 are alsoa function of the proper pipe size andlength for pipes 1, 2 and 3, and theircorresponding friction losses (F1, F2,F3). The suction side friction lossesare assumed negligible in this analy-sis. Clearly, source pressures, if aboveatmospheric, must be included.

There are three possible operat-ing conditions—(A) both pumpsoperating or off, (B) #1 on and #2 off,and (C) #1 off and #2 on.

Calculating the pump dischargepressures, P1 and P2, at given flowrates, Q1 and Q2, is the first step insizing a pump. Variations in pressureand or flow rate provide the pump(s)operating head range, important cri-teria in the final selection.

Using the information shown inFigure 1, we will solve for the pres-sures P2 and P1. For example, thepressure at P2 will be the algebraicsum of destination pressure (P3),friction losses (F3 and F2) and thedifferential elevation (Z3–Z2), allexpressed in the same units (ft of liq-uid flowing or psi). This can be

expressed as follows:

P2=P3+F3+F2+(Z3–Z2)P1=P3+F3+F1+(Z3–Z1)

F3 calculated at Q1+Q2 flow

For the three conditions

A (both #1 and #2 pumps oper-ating) use the expressions above

B (#1 pump on, #2 off)

P2=P3+F3+(Z3–Z2)P1=P3+F3+F1+(Z3–Z1)

F3 calculated at Q1 flow

C (#1 pump off, #2 on)

P2=P3+F3+F2+(Z3–Z2)P1=P3+F3+(Z3–Z1)

F3 calculated at Q2 flow

Calculating the pressures for thethree conditions gives the operatingpressure ranges for pumps 1 and 2.Normally, the pipe lengths and flowrates (or flow ranges) are fixed by thesystem design. The engineer canadjust pump design pressures bychanging pipe size. This can simplifypump selection by reducing thepump(s) operating head range for thevarious operating conditions.

In Figure 1 we have listed thepipe lengths and Q1 and Q2 setflows. Consider Z1, Z2 and Z3 allzero. (There is just enough pressureto get liquid to pumps 1 and 2 attheir set flow rates.) Why not deter-mine the optimum Schedule 40 pipesize to minimize the variation inpump pressure for the three condi-tions of operation?

If you have an expert program,you can verify your manual calcula-tions. Everybody seems to have accessto the Cameron Hydraulic Data book,so we can use its friction loss numbersfor new pipe. I will discuss the resultsbriefly in next month’s column, whenwe will also look at another, morecomplex problem. ■

Until next time…

Pump 1Q1

FIGURE 1. Multi-source single destination system

Pipe 1-300 ft Q1 = 200 GPMPipe 2-100 ft Q2 = 50 GPMPipe 3-300 ft P3 = 30 PSI (69.3ft)Liquid water at 60°F

21 3

P3Z3

P2

P1

Z2

Z1

Pump 2Q2


HANDBOOK??


Sizing Pumps for Complex Systems: Part IIBy J. Robert Krebs, P.E., Contributing Editor

Continuing the multi-sourcesingle-destination discussionbegun last month, we willstudy the problem of four

pump stations pumping to a singledestination system as shown in Figure 1. Going from a two sourceto a four source single-destination system greatly complicates thedetermination of individual pumpdischarge pressures when one ormore pump stations (PS) are notoperating.

The system illustrated is a com-mon problem in wastewater pump-ing. For example, perhaps stations 2,3 and 4 were added to the originalPS1 with discharge line pipes 1, 3, 5and 7. The addition of each PSaffects the operation of the PSalready on line and may requireequipment or piping changes witheach additional PS.

The method used in last month’scolumn will again be employed tocalculate the pump discharge pres-

sures manually for the four sourcesingle destination system.

Recall that in the two-source,single-destination system, there weretwo basic equations to calculate thedischarge pressure of the two pumpsources. With a four source single-destination system, there are fourequations, one for each pump dis-charge pressure.

EQN1 P1 = P5+F7+F5+F3+F1+(Z5-Z1)

EQN2 P2 = P5+F7+F5+F3+F2+(Z5-Z2)*

EQN3 P3 = P5+F7+F5+F4+(Z5-Z3)*

EQN4 P4 = P5+F7+F6+(Z5-Z4)*

* The pump discharge pressureis to be corrected to pump centerline.

An important consideration—the friction loss in any line is calcu-lated as the sum of the flows in thatline. For instance, F7 would be cal-culated as the sum of flowsQ1+Q2+Q3+Q4 if all stations wereoperating, and for the same condi-tion F3 would be calculated as thesum of flows Q1 and Q2.

As an example, consider pump2 out of service and pumps 1, 3 and4 operating. The friction losses F3,F5 and F7 would be calculatedwithout the flow Q2, and the pres-sure at P2 would be the pressure atthe intersection of pipes 2 and 3(P5+F7+F5+F3+(Z5-Z2) with Z2corrected to pump 2 centerline.

It is obvious that to balance theflow and size pipes for this exampleis much more complicated than forthe two source system. I use a sim-ple expert pipe network programfor this type of problem. When Ientered into the computer the flowvalues, pipe sizes and lengths


P1

Pipes 2 & 4 = 100 ft Q1 = 50 GPMPipes 1, 3 & 6 = 200 ft Q2 = 100 GPMPipe 5 = 400 ft Q3 = 200 GPMPipe 7 = 500 ft Q4 = 300 GPMLiquid Water 60°F P5 = 69.3 ft(2.5) = Pipe Size Inches

P2

P3

P4

P5Z5

Z4

Z3

Z2

Z1

PUMP 2

PUMP 1

PUMP 3

PUMP 4

1

2

3

4

5

6

7(2)

(2)

(2.5)

(3)

(6)

(4)

(6)


HANDBOOK??


shown in this example, I saw thatthe results using four pump curvesI selected closely matched my man-ual calculations using new pipe fric-tion factors.

Try it! If you have any questions,let me hear from you.

The last two columns havefocused on open transfer systems.Next month we will try a more com-plex closed system.

Incidentally, I would like tothank the many readers who calledand wrote for more details on speci-fying shaft deflection and bearinglife after reading the May article. ■

Until next time ...

JULY ISSUE—PART I SOLUTION

Using the two equations for man-ual solution of the pump dischargepressure requires pipe size selectionfor the three pipes. To calculate thefriction losses, Figure 1 gives the pipesizes I chose. With pump flow set atthe given value, if pipes 1 and 2 aresized so that their friction losses areapproximately equal, the effect ontotal head with either pump out ofservice will be minimal.

With the pump curves used (seeFigure 1B), the effect on head ofeither pump working versus bothworking was 2-4%. Capacity varied50-60 gpm and 230-237 gpm for

pumps 2 and 1 respectively.In the actual system design, cen-

trifugal pump performance curveswould be used. Pump curves withsimilar characteristics should be cho-sen. For example, if a pump is select-ed that has a drop in total head of50% from shutoff (zero flow) to bestefficiency flow, then other pumpsshould be similarly selected. Also, itis better to select the larger flowpumps before smaller flow units. Thepump curves I chose show the effectof one and two pump operation.

PUMP 1

PUMP 2

Flow Rate GPM0 100 200 300 400

Head

160 ft.

120 ft.

80 ft.

40 ft.

0 ft.

FIGURE 1B. Pump curves

PUMP CURVES

Pump 1Q1


Pipe 1-300 ft Q1 = 200 GPMPipe 2-100 ft Q2 = 50 GPMPipe 3-300 ft P3 = 30 PSI (69.3ft)Liquid water at 60°F

1 - 300 ft - 3.5 INP3

Z3

P2

P1

Z2

Z1

Pump 2Q2 2 - 100 ft - 1.5 IN

3-300 ft - 4.0 IN


Continuing the discussion oncomplex systems in this col-umn, we will examine closedsystems. Recall there are two

basic pumping system types – open ortransfer systems and closed or circu-lating systems. Either type may havesingle or multiple sources and singleor multiple liquid destinations. Theleast complicated closed systemwould be a single source pumping to asingle destination and returning all ofthe pumpage to the source. An exam-ple might be pumping liquid from acontrolled temperature source to adestination where some of the liquidthermal energy is absorbed or dissi-pated and the liquid is then returnedto the source. The liquids used in suchsystems are often specially formulatedheat transfer liquids. One well knownapplication is chilled water systems incommercial buildings and similarequipment used in industrial settings.

Adding two destinations willmake the closed circulating systemjust described more complex. Figure 1illustrates such a single source multi-destination closed or circulating pumpsystem.

We will consider the destinationsA, B and C identical. They could bebatch process or continuous flow ves-sels, each with a flow control valve(FCV) to regulate the flow rate to theprocess energy requirement. A reliefvalve (RV) bypasses excess flow at aset pressure. I realize other accessoriesare needed to complete the system.However, the system in Figure 1 willserve our purpose for this discussion.

The question is how to size apump (or pumps) to provide a correctsupply of liquid. The designer wouldstart with the maximum energyrequirement (BTU/hr), the thermalcapacity of the liquid, the allowabletemperature gradient to develop flowrate. Next is the question of how tocalculate the pressure required for thepump(s). The heat transfer liquidmanufacturer would provide the liq-uid physical properties and pipe pres-sure loss charts to facilitate the totalhead calculation. To simplify the sys-

tem calculations, if you wish to try asample problem, assume the liquid iswater.

Remember that in these systemdesigns our objective is to establish amethod, if possible, to calculate thetotal head requirement manually. Asthe systems become more complex,network software is the only way toplay “what if” and establish theboundary conditions quickly.

Referring to the figure, assumethat only destination C is receiving liq-uid. The pressure gauge P1 wouldthen read the sum of the pressuredrops from pipe friction through thesystem at the flow rate of the pump,which will include bypass flow pres-sure from the relief valve (RV) or

P1=F11+F10+F9+F8+F7+F6+F3+F2+F1

for only destination C operable.

No elevation head differentialsare considered.

The same formula arrangementwould apply to any single destinationoperation by incorporating the appro-priate pipes.

For a manual calculation for two,three (or more) destinations that areidentical, I have found that adding theboundary friction losses and takingthe average of the destination lossesclosely approximates the results froman expert network software program.

If all three identical destinations

were operating, the pump gauge pres-sure would read the sum of theboundary friction losses plus the aver-age losses across the three destina-tions plus any pressure drop across anoperating relief valve.

P1=F11+F10+F9+F8+F7+[F6+F5+F4]/3+F3+F2+F1

for three identical destinationsoperating.

If the destinations are substantial-ly different in friction loss, using themaximum loss unit would producethe greatest head requirement.

I hear a chorus of voices rising,saying that this is an excellent appli-cation for a variable speed pump,and it is.

With constant speed operation,the RV would be bypassing substan-tial flows with only one destinationactive or with one or more destina-tions operating at less than maximumflow. The pump is operating at a con-stant flow and pressure or a fixedhorsepower.

A variable speed pumping sys-tem, with the motor speed controlsensing the total FCV requirement,would permit the pump to operateover a speed range based on demand,and it would decrease the pump pow-er required. ■

Until next time ...

Sizing Pumps for Complex Systems: Part IIIBy J. Robert Krebs, P.E., Contributing Editor

FIGURE 1. Single source multi-destination closed system

10 9 8 7

654

321

11

P1

RV A B C

FCV

Q


HANDBOOK??


Air is used as the motive forcein air motor driven pumps,some forms of diaphragmpumps, pneumatic ejectors

(which are enjoying an increasedpopularity for laboratory, blending,drainage and sampling systems) andin air lift pumps. The special air liftpump we will discuss in this month’scolumn has been around a long time.Some of the earliest designs were forpumping water from wells. Figure 1shows the general design.

A vertical pipe extends into a liq-uid reservoir. The pipe is fitted near

the bottom with a diffuser mecha-nism that will let a compressed gas(air) into the liquid column.

The gas bubbles will risethrough the static liquid column inthe pipe, displacing liquid. Theresulting gas-liquid mixture has aspecific gravity (specific weight) lessthan that of the liquid. The liquid-gasmixture rises in the pipe in propor-tion to the submergence of the pipein the reservoir. When the reservoirsubmergence head exceeds the staticlift and pipe friction losses, the liquidis discharged from the system and

the spent gas is wasted. The energylevel of the discharged liquid is thepotential energy of its elevationabove some destination to which theflow may be directed by gravity.

In the most common applicationof lifting water, the gas would beair—supplied at an appropriate pres-sure and quantity to pump thedesired volume of liquid at thedesired rate.

Looking at the air lift pumpdesign, it is clear that there is theadvantage of no moving parts. Also,the air is compressed at a location farremoved from the pumping action,where equipment can be protectedas required.

Liquids containing solids, slur-ries and abrasive solids such as sandor mine tailings move freelythrough the pump without prob-lems. The simple design can bemade in almost any shop. Laborato-ry and process applications are aninteresting area where the pumpmight be used. Common applica-tions in smaller waste treatmentplants include pumping return andwaste activated sludge. In digestersthe mixing of biological anaerobicsludge can be accomplished with aspecial air lift pump, where the gasused is methane under pressure.The methane is produced by theprocess. Sequencing batch reactors(SBR) and other treatment processesalso have applications. Some addi-tional applications include drainageof wetlands and the obvious trans-fer of water to a discharge point.Note that the applications men-tioned all involve low heads and rel-atively high volumes.

In matching the product to theapplication, one must consider prod-uct and system constraints. In myopinion, the best application area isfor relatively high flows and lowheads, where the efficiency and con-

Pumping with AirBy J. Robert Krebs, P.E., Contributing Editor

FIGURE 1. Air lift pump

AIR

DISCHARGEDLIQUID

AIR UNDERPRESSURE

STATIC LIFT

SUBMERGENCE

AIRDIFFUSER


HANDBOOK??


trol is best. Also, the solubility of thegas in the liquid must be considered.This could be a plus or minus in var-ious lab and process applications.

Working with static lifts of 4-6feet with 65-80 percent submergencewill give a pump design of reason-able efficiency. Pipe friction lossesthat negate the submergence effectshould be minimized.

Something should be said of thedisadvantages of the air lift principle.Flow control can be difficult. Thepump percent submergence designand head, and the air pressure anddiffuser design, will determine howwell the flow can be controlled,though varying the flow accuratelywill still be a problem.

The gas bubble size is dependenton the dispersing method. The mostcommon design is a series of holesaround the periphery of the pipewith a collar covering the holes andthe air pipe connected to the collar as

shown in Figure 1. Bubble size isimportant because the bubble riserate increases with bubble size, asdoes the drag force on bubbles risingthrough the liquid. As the liquid-gascolumn accelerates to some steadystate condition, the relative motionbetween the bubbles and the risingwater column will decrease remark-ably. The bubbles’ internal pressurewill decrease as the liquid-gas mix-ture rises. This will increase the bub-ble diameter and the mixturevelocity. This natural action will helpexplain the problem of flow control.

This is a brief review of a usefulproduct.

The literature sources I checkedpresented minimal information onpipe sizing, flow rates and pressure.It seems that the equipment manu-facturers that make pumps of thistype for their own use are the custo-dians of the design information.

From the limited data found, I

was able to formulate an application,sizing and design approach. Before Ipublish this approach, I would like toconfirm the theory against practice. Iwould ask readers of this column topass along their design criteria. Ifyou do not want your name men-tioned when I acknowledge yourcontribution, please say so.

In return, I will attempt to con-solidate the data to provide a generaldesign form for air lift pumps. ■

Until next time...

To send or fax informationregarding the design criteria;

Pumps and Systems MagazineAttn: Bob Krebs123 N. College Ave., Suite 260, Ft. Collins, CO 80524Fax (907) 221-2019


Effective troubleshooting shouldbe an integral part of anyplant’s equipment reliabilityand maintenance program.

Before you examine the system, how-ever, take time to review maintenancerecords. The performance history ofyour pumps will point you in the rightdirection. Also, pump operators canprovide clues to what is happening.Troubleshooting should focus onthese areas of investigation:

1. THE FOUNDATIONPoor foundations, grouting and

baseplate design often cause prob-lems.

2. DRIVERVibrations of the driver (motor,

steam turbine, gearing) can be trans-mitted to other components.

3. MECHANICAL POWERTRANSMISSION

Excitations from the couplingarea, especially because of driver mis-alignment or eccentrically bored cou-

pling hubs, can also be transmitted toother parts of the pump and/or sys-tem. Users should note any incorrectpositioning of the driver and pumpsuch that the distance between shaftends (DBSE) exceeds the axial flexinglimits of the coupling.

4. THE DRIVEN PUMPPump design has a major influ-

ence on the hydraulic interactionbetween the rotor and the casingand consequently the problemsencountered. Misconceptions aboutpump thermal-growth can createproblems.

5. SUCTION PIPING AND VALVESImproper design and layout of suc-

tion piping and valves can create flowdisturbances such as cavitation, intakevortexing or suction recirculation.

6. DISCHARGE PIPING AND VALVESUnfavorable dynamic behavior

of piping resulting from loads trace-able to dynamic, static or thermalcauses (including resonance excita-

tion) can create trouble.

7. INSTRUMENTATION FOR CONTROL OF PUMP FLOW

Pressure pulsations can resultfrom control system-pump interac-tion during start-ups, periods of lowflow and valve changes.

8. MAINTAINING ALIGNMENTOnce the alignment is estab-

lished, dowels into the baseplatemust hold the pump in alignment.

Troubleshooting centrifugalpumps begins with observing operat-ing conditions at the site. While amyriad of problems can exist withany pumping system, here are someof the more typical scenarios, possi-ble causes and corrective actions. ■

Ed Nelson has more than 40 yearsof experience with industrial pumpingsystems as a former end-user at a majorpetrochemical company, and today as aturbomachinery consultant. He is amember of the Pumps and SystemsEditorial Advisory Board.

Troubleshooting Centrifugal Pumps

Cavitating-type ProblemsA cavitating sound is heard in a pump that does not normally cavitate, and it is not clear

whether it is pumping into the system.

POSSIBLE CAUSES CORRECTIVE ACTION

The suction piping layout is poor. There may be too Redesign piping layout, using fewer ells and laterals formany ells in too many planes or not enough straight tees, and have five or more diameters of straight pipe run before the suction flange of the pump. before suction flange.

Suction piping configuration causes fluid to rotate Install enough straight run of suction piping, oradversely when approaching impeller. install vanes in piping to break up prerotation.

Flow rate is high enough above design that NPSHr Reduce flow rate to the level for which the pumphas increased above NPSHa. was designed.

Pump is operating at a low-flow, producing suction Install bypass piping back tosuction vessel to increase recirculation in the impeller eye, resulting in a flow through pump. Note: Bypass flow may have to be cavitation-like sound as high as 50% of design flow.

The suction screen is clogged. If screen is present, remove and clean it.

Piping gaskets with undersized IDs have been Install properly sized gaskets. installed—a common problem in installations of small pumps.

Pipelines are constricted because of buildup of Replace deteriorated pipe.corrosion materials.


HANDBOOK?


Capacity-type Problems Pump does not have enough capacity. No significant noise is coming from it.

POSSIBLE CAUSES CORRECTIVE ACTIONPump suction is below atmospheric pressure, Eliminate air leaks with appropriate actions.causing air leaks into system.

Wear ring clearances are excessive (closed Overhaul pump. Replace wear rings if clearance is impeller design). about twice design value for energy and performance

reasons.

Impeller-to-case or head clearances are Reposition impeller for correct clearance.excessive (open impeller design).

The discharge block valve is partially closed. Open valve completely.

Any of the following conditions may Do the following: have increased friction in the piping to the discharge vessel:1. Gate has fallen off the discharge valve stem. 1. Repair or replace gate valve.2. Check valve spring is broken. 2. Repair valve by replacing spring.3. Check valve flapper pin is worn, and the 3. Overhaul check valve. Restore proper clearanceflapper will not swing open. to pin and flapper bore.4. There is collapse of lined pipe. 4. Replace damaged pipe.5. The control valve stroke is improperly set, 5. Adjust control valve stroke as needed.resulting in too much pressure drop.

Suction and/or discharge vessel levels are not correct. Calibrate level controllers as needed.

Motor is running backward or impeller of double Check for proper rotation and mounting of impeller. suction pump is mounted backward. Discharge pressure Reverse motor leads if necessary. developed in both cases is about one-half design value.

Entrained gas from the process is lowering NPSH Reduce entrained gas in liquid by process changes asavailable. needed.

Mechanical seal in suction system under vacuum is Change percentage balance of seal faces or increase leaking air into system, causing pump curve to drop. spring tension.

There is polymer or scale buildup in discharge Shut down pump and remove scale or deposits.nozzle areas.

Vortex formed at high flow rates or low liquid level. Reduce flow to design rates. Raise liquid level in Does the vessel have a vortex breaker? Does incoming suction vessel. Install vortex breaker in suction vessel.flow cause surface to swirl or agitate?

A variable speed motor is operating too slowly. Adjust motor speed as needed.

Bypassing is occurring between volute channels in a Overhaul pump. Repair damaged areas. double volute pump casing due to a casting defect or extreme erosion.

Axial positions of impeller(s) are not centered with Overhaul pump; reposition individual impellers as diffuser vanes. Offset of several impellers will cause needed. Reposition rotor by changing thrust collar vibration and lower head output. locator spacer.

When the suction system is under-vacuumed, Install a positive-pressure steam (from running the spare pump has difficulty getting into system. pump) to fill the suction line from the block valve

through the check valve.

Some pump designs incorporate an internal bypass Overhaul pump. Restore orifice to correct size. orifice port to change head-flow curve. However, high liquid velocities often erode the orifice, causing the pump to go farther out on the pump curve. The system head curve increase corrects the flow back up the curve.

The volute and cutwater area of casing is Overhaul pump. Replace casing or repair by welding. severely eroded. Stress- relieve after welding as needed.

A replacement impeller does not have a correct Overhaul pump. Replace impeller with correct pattern.casting pattern, so NPSH required is different.


Motor Overload ProblemsPOSSIBLE CAUSES CORRECTIVE ACTIONPump is circulating excessive liquid Overhaul pump, replacing parts back to suctionthrough a breakdown needed. bushing or a diffuser gasket area.

There is polymer buildup between wear surfaces Remove buildup to restore clearances.(rings or vanes).

There is excessive wear ring (closed impeller) or Replace wear rings or adjust axial clearance of open cover-case clearance (open impeller). impeller. In severe situations cover or case must be

replaced.

Open impeller has slight rub on casing. This Increase clearance of impeller to casing. usually occurs in operations from 250–400°F due to piping strain and differential growth in the pump.

The minimum flow loop has been inadvertently left open Close minimum flow loop or control valve bypass valve. at normal rates, or bypass around control valve is open.

Discharge piping is leaking beneath liquid level Inspect piping for leakage. Replace as needed.in sump-type design.

One phase has low amperage due to electrical switch Check out switch gear and repair as necessary. gear problems.

Specific gravity is higher than design specification. Change process to adjust specific gravity to design value, or throttle pump to reduce horsepower require-ments. (Note: This will not correct the problem withsome vertical turbine pumps which have a flat horse-power required curve.)

Pump motor not sized for end-of-curve operation. Replace motor with a larger size, or reduce flow rate.

A replacement impeller was not trimmed to the Remove impeller from pump. Turn to correct diameter.correct diameter.


In the November1995 and January1996 issues this column identifiedthe cause and effect of water ham-mer in piping systems. A gravity

flow and a pumped system wereused to illustrate conditions con-ducive to the development of poten-tially dangerous pipe surge pressuresand to show how the maximum pres-sures and cycle time can be calculat-ed. Twelve surge signals were alsoprovided to serve as a guide in deter-mining the need for a closer look atsurge pressure potential. Both the

formulas and the surge signals arereproduced in an accompanying pre-sentation this month.

It should again be stressed thatthe rigid pipe theory approach andthe 12 surge signals provide a quickcheck to judge the need for a morecomplete elastic pipe theory analy-sis. They are not a substitute for themore detailed approach.

Figures 1 and 2 illustrate gravityand pumped flow systems respec-tively with the piping design thesame in each.

The values of the cycle timeperiod, maximum surge pressureand pipe design criteria are listed ineach figure. One should first examinethe surge signals and decide which,if any, apply to the problem. Thenselect the method of surge attenua-tion and the size of the device(s), ifneeded, or proceed to a completeelastic pipe analysis.

For Figure 1, items 2, 3, 4 and 6of the surge signals are applicable.Power outages (item 2) are the over-riding rationale for most surge pro-tection devices used. Normally,emergency power systems cannotrespond quickly enough to avoidactivation of a pressure surge. Valveclose (and opening) time (item 3) canbe controlled in the system design.This can reduce, even eliminate, thesurge pressure in normal cycle oper-ation. If pneumatic or hydraulicvalve actuators are used, a pressuredfluid supply (enough for severalvalve operations) can reduce or elim-inate the danger from power outages.Pipe and accessory equipmentdesign pressures (item 4) with surgeallowance must be higher than thecalculated total pressure includingsurge. Note that if repeated surgesare possible, the repeated cumulativestresses could produce a fatigue fail-ure. The endurance limit of materi-als, including plastics, must beconsidered. The use of stronger pip-ing and equipment can be cost effec-tive compared to surge suppressionequipment.

Items 4 and 6 can be consideredtogether. Calculated results for maxi-mum pressure and cycle time periodare shown. The calculated maximumpressure exceeds the pipe designpressure with surge allowance.Increasing the valve closing time isan option. If the on-off valve has alinear closing characteristic andcould be closed uniformly, in say 15sec, to reflect a 75% drop in flow (or,say, 2 ft/sec followed by a 5 sec inter-val to close), the maximum surge

Water Hammer – Containing the SurgeBy J. Robert Krebs, P.E., Contributing Editor


HANDBOOK??


Figure 1. Gravity flow system – surge relief

Reservoir

Reservoir

On-off valve

SRV

Figure 2. Pumped flow system – surge relief

Source

DestinationSRV

HS=100 ft

Q=4000gpm, V=6.4fps, VC=4000fps, T=6.7secH=795 ft + 100 ft = 895 ft (387psi)

2.5 MI (13333 ft)16" ID pipe150psi Design Plus100psi Surge Allow

100 ft

Pipe Same as Figure 1T = 6.7secH = 795 ft + 200 ft = 995 ft (431psi)

200

100

00 4000

Q GPM

HeadFT

CCV

CCV

MaxRPM

REDRPM

pressure would be significantlyreduced to about 150 psi.

For the Figure 1 gravity flow sys-tem, regulating the valve closuretime could satisfy all concerns exceptpower outage. A surge relief valve isthe obvious answer for that concernand would most likely be less expen-sive than changing the pipe designspecifications.

The surge relief valve (SRV)pressure setting would allow the SRVto open as necessary on power out-ages. Two manufacturers recom-mend either a 6- or 8-in SRV for thisapplication.

The Figure 2 pumped systemwould be compared with the 12surge signals. Items 2, 3, 4 and 6 areagain not in agreement with thedesign. A power outage (item 2)would interrupt flow abruptly as thepump or pumps stopped requiringsome form of system surge pressureprotection. A controlled closingcheck valve (CCV in Figure 2) wouldpermit an extended closing time(item 3) as in Figure 1, when a con-trol valve was substituted for the on-off valve. As the valve closes, theadditional pressure drop across thevalve reduces the flow rate. At somepreset time or valve position, thevalve closes and the pump stops. Asimilar effect can be accomplishedby using a variable speed driver andreducing speed to reduce head andflow prior to shutoff of this pump.Both methods are shown on thepump and system curve in Figure 2.The power outage concern can becovered by an SRV as shown in theFigure 1 solution. A surge reliefvalve would be sized at 10 in by onemanufacturer.

Higher design strength (item 4)pipe and fittings could be consid-ered. However, another considera-tion (item 6) of pipe velocity and theaccompanying friction loss should beconsidered. If 18" ID pipe were con-sidered, the velocity at 4000 gpmcomes down to 5.04 ft/sec, reducingthe friction loss to 56 ft and the max-imum surge pressure to 626 ft or atotal pressure of 682 ft (295 psi). Thelower friction loss will also bereflected in smaller driver 232 BHPvs. 190 BHP and a substantial powersavings of $10.44 per million gallons

pumped (at $0.08 per KwH).To summarize the Figure 2 solu-

tion, I would consider using the 18"pipe with the 150 design and 100 psisurge allowance. While a controlvalve or controlled closing checkvalve and constant speed drives arepossibly a less expensive solution, Iwould prefer using a swing checkvalve (perhaps cushioned) and a vari-able speed driver (VSD). The controlvalve method of reducing flowimposes higher pressures on thepump, as is obvious from the pumpsystem curve. Conversely the flowrate and pressure are reduced asspeed is reduced with the VSD.There is no single protective deviceto solve all water hammer problems;rather, there are many devices thatcan absorb or eliminate dangeroussurge pressures. To name some thatare regularly used – in line checkvalves for low head systems, bladderand hydro pneumatic tanks (Helm-holtz resonators) and surge towers.There are few designs that are asstraightforward – simple might bemore correct – as these two exam-ples. These articles have been in-tended to provide an introductoryexplanation of possible solutions towater hammer problems in a system.

Finally, system operating condi-tions are all important. Initial fillingof the piping system must be consid-ered with the operation of vent andair release valves. Oversized surgerelief valves can promote dangerouspressure surges, strengthening thecase for multiple-sized valves withseparate pressure settings in largersystems. ■

Until next time ...

Maximum SurgePressure – Cycle Time

CalculationsH = Vc x V/g + HsVc = Sonic velocity ft/sec

V = Steady state liquid velocity, at flow interruption

g = Gravitational constantft/sec2

Hs = Static or operating head ft

H = Maximum pressure in ft of liquid flowing

T = 2L/Vc L = Pipe length ftT = Cycle time in sec

Twelve Surge Signals1. Pipe or fittings broken or

cracked (existing system)2. Power outage – emergency

power normally not a solution3. Valve closes in less than cycle

time period4. Maximum calculated surge

pressure plus static or operatinghead approaches the pipe designincluding surge allowance

5. Large elevation changes (34 ftor one bar) over short pipe lengths

6. Higher pipe line velocities,say, above 5 ft/sec

7. System renovation with greaterflow through an existing system

8. A low head system that per-mits continued flow through a pumpafter shutdown

9. Dead ends in piped systemsthat can promote higher than calcu-lated pressure surges through reso-nant conditions

10. An undulating pipe linetopography with lower head encour-aging vapor cavity formation

11. Long suction lines forpumped systems

12. Improperly sized vent andvacuum valves

Pumping With Air

Many readers responded withtheir information on air lift pumps.To the ten of you who wrote and themany of you who called my officewith their helpful contributions asincere thank you! From my posi-tion, putting it all together is not asstraight forward as I had suspected.As soon as it is completed, each per-son who provided information willhave the results and a column will bewritten. Thanks again!!



HANDBOOK??


Reducing Noise In API Process PumpsTests and observations point to geometry change as solution.By Leon K. Stanmore

In offshore drilling operations, APIprocess pumps sometimes have tobe mounted on structural steel

platforms. Because the pump assem-bly is not attached to a concrete foun-dation and the baseplate is not filledwith grout for rigidity and vibrationdamping, the movement and noisegenerated by a centrifugal pump areof special concern. Steel structuressupporting rotating equipment notonly transmit but amplify even thesmallest vibrations or noise.

In one such application it wasdetermined that the maximum vibra-tion had to be kept below 0.10 ips andthat no discernible hydraulic noisewould be permitted. Two 8x10x19horizontal API 610 process pumpswere brought in from the field and setup on a test floor to investigate proce-dures for minimizing random noise.

The units were identical top suc-tion and top discharge API 610 cen-trifugal pumps with overhung singleentry impellers.

The pumps are normallyinstalled on an offshore platform.Both have ample NPSHR. The pumpsupporting base is made of weldedstructural steel, and there is no groutto provide damping or rigidity. Endsuction, top discharge pumps havebeen used successfully in this situa-tion since 1956. However, top suctiondesigns were added later, and fewpumps with the bell-mouth designwere built and tested.

The pumps in question, tested atduty conditions of Q=2150 gpm,H=314 ft and N=1770 rpm, generat-ed intermittent bursts of broadbandnoise. Reviews of test logs did notindicate any unusual pump behavior.

PROBLEM DISCUSSIONWhen pressure fluctuations are

produced by liquid motion, thesources are fluid dynamic in charac-ter. Fluid dynamic sources include

“X” “Y”

“C”

Figure 1.

Figure 2.

Figure 1. Before geometry changeFigure 2. After geometry change


turbulence, flow separation, cavita-tion, water hammer, flashing andimpeller interaction with the case atthe water or inlet guide vane (split-ter).

The pressure and flow pulsationmay be periodic or broadband in fre-quency and excite either the pipingor the pump into mechanical vibra-tion. Mechanical vibrations can thenradiate acoustic noise into the envi-ronment.

Pump pulsating sources are clas-sified as:

• discrete frequency componentsproduced by the pump impeller

• broadband turbulent energyresulting from areas of high veloci-ties

• impact noise consisting ofintermittent bursts of broadbandnoise caused by cavitation, flashing,water shearing action, water ham-mer and pockets of entrapped air atthe suction inlet near the impeller

• flow-induced pulsation causedby periodic vortex formation whenthe pump fluid passes an obstruction.

Secondary flow patterns that canproduce pressure fluctuations in cen-trifugal pumps include stall, recircu-lation (secondary flow), circulation,leakage, cavitation and wake (vor-tices) at inlet guide vane (splitter).

PROCEDURESOne of the pumps was opened

and examined for any possiblerestrictions that could trap air. Thepump case and impeller were exam-ined for dimensional compliance,casting quality, machining, balancingand cavitation or mechanical dam-age.

During the investigation it wasdiscovered that the suction flow split-ter had square edges instead of radiifacing and trailing the flow. Thiswould create a disturbed flow whenin contact with the square edge split-ter. Furthermore, the square edge ofthe 6-3/4"splitter (edge X in Figure 1)would create turbulence as well as achange of momentum. If we considerthe splitter as a stationary blade, thena careful examination of the splitterin the suction bell-mouth led us tobelieve that the splitter design and itsposition relative to the impeller inletwere not optimized. It was felt thatthe boundary layer on the splitter,when reaching the square edge of the

splitter "separation" points, causedthe wake, together with thefreestream, to extend into the im-peller eye.

Greater distance between thesplitter's vertical edge and theimpeller inlet was considered impor-tant. The impeller outlet angle wasβ=24º, and the francis vane inletangles were shroud 17º, hub 30º.There were 7 vanes, and the gapbetween the impeller's outside diam-eter and the case cut water wasexpressed as:

B=100 (R3-R2) =12%R2

hence, this was consideredacceptable for a volute type pump.

The impeller eye was reducedfrom a diameter of 9 1/2" to 8 5/8",thus raising the velocity of fluid mov-ing through the eye from 10.76 fps to13.37 fps. Subsequent tests indicatedthat intermittent noise was reducedby 50%. Both pumps were the cus-tomer's property and a little late ondelivery, so we did not have the timeto install windows in the suction bell-mouth for visual observation andstudy. Again, we examined the split-ter in the section bell-mouth anddetermined that its design and itsposition relative to the impellervanes at inlet were not optimized.

The distance between the split-ter edge (Figure 1, "Y") and theimpeller eye was considered insuffi-

cient to accommodate the unevenflow. The area marked "C" in Figure1 is a pocket in the case castingwhere the velocity is very low as theflow changes its direction by 90º.This area was considered a possibletrap for entrained air. The observednoise could be best described as ran-dom "shearing" action of water.

The pump was tested fromclosed valve to open valve flow. Seecurve 4590 for flow, head efficiencyand NPSHR. Careful observationswere made as to the noise intensity.With the flow reduced to 1,150 gpm,the noise was slightly less but stillnot acceptable.

RESULTS OF GEOMETRY CHANGESIt was reasoned that the splitter

required a geometry change fromFigure 1 to Figure 2. All edges weretherefore rounded, and the pumpwas retested. The intermittenthydraulic noise disappeared. TheNPSHR went up 1.6 ft with the 8 5/8" eye ring. Flow head character-istics were unaffected, the H-Q curvewas stable, and there was no changein pump efficiency. Vibration in thevertical plane, as measured at theradial pump bearing, went up from0.04 to 0.07 ips at the speed of rota-tion, 1770 rpm.

RECOMMENDATIONS ANDOBSERVATIONS

Based on the observations and

400

350

300

250

200

100

90

80

70

60

50

40

400

200

00 500 1000 1500 2000 2500 3000 3500

Tota

l Hea

d in

Fee

t

Horsepower

% Efficiency

NPSH in Feet

Gallons Per Minute

TDH VS CAP

EFF

40

20

0

NPSH Req'd

BHP @ SP GB = 822

Figure 3. Performance curve of 8x10x19, 1770 rpm API process pump


data collected in this application, it isrecommended in general that anychanges in the suction inlet areawhere a splitter is located should beevaluated first by checking the veloc-ity flow, cascade and geometry. Thisshould be followed by comprehensivetests at all flows in the range of appli-cation. It is essential that uniformvelocities are present from the suc-tion flange through the passageswhere the flow will accelerate intothe impeller eye.

It appears that other pump sizeswith the same splitter configurationworked fine, but in our case theresults were unacceptable. From this,it is clear that it is necessary to verifyany geometry changes by full speedhydraulic tests at all flow conditions.

If the pump had been fitted withPlexiglas windows in the area of thesuction splitter and the impellervanes at inlet, it would have beenpossible to observe the separation offlow at the splitter, as well as thewake extending into the impeller eye.One also might have seen entrainedair in the low velocity area, "C" in Fig-ure 1. Generally, though, a full rangeof flow testing can provide the sameresults if it is conducted thoroughly.■

Leon K. Stanmore is a centrifugalpump consultant with more than 40years of experience in research, design,development and analysis. He is pastchairman of the Reciprocating PumpSection of the Hydraulic Institute and atask force member of API Standard 685on Sealless Centrifugal Pumps.

The boiler feedwater system atthe Novacor Chemicals world-scale ethylene plant in Sarnia,Ontario included horizontal

multistage barrel case pumps and sin-gle stage impulse steam turbine dri-vers with mechanical governors. Aspart of a major effort to eliminate aproduction bottleneck, the pumps anddrivers were rerated to provide moreflow. The entire pump inner caseassemblies (including rotor and innercasing) were replaced with highercapacity elements. New governorvalves and nozzle rings were installedon the turbines to produce more pow-er. And an electronic governor thatadjusted the pump speed to maintainconstant header pressure was added tothe control scheme.

The new equipment was in-stalled in September 1994. Nothingwent well. The pump suffered highblade pass frequency vibrations. Thesolution required unique hydraulicmodifications and inboard bearingbracket stiffening to eliminate reso-nance. The turbines tripped mysteri-ously and broke governor valve stems.A valve resonance problem was solvedby installing a valve with a differentgeometry. Teething problems with theelectronic governor control systemalso had to be resolved. These prob-lems became very high profile whenthe newly rerated boiler feedwaterpumps tripped and caused a total plantshutdown resulting in a 14-hour flare.Not only did lost production reducerevenues, but the Ministry of the Envi-ronment initiated a review of ourentire flaring history.

Although it took a year of difficulteffort, the pumps and turbines arenow operating satisfactorily. Followingis a chronology of these problemevents and the resulting solutions.

BACKGROUNDThe boiler feedwater pumps and

turbines had operated reliably for thepast 18 years. Overall vibration levelswere around 0.2 in/sec. (All vibrationreadings have units of inches per sec-ond, zero to peak, calculated and were

taken with magnetic base accelerome-ters and a portable dual channel ana-lyzer using a Hanning window). Wenormally operate two steam turbinedriven pumps (designated A and B).An electric driven spare pumpremains on hot standby but we avoidoperating it because it results in highdemand charges from the power com-pany. If any pump is unavailable theOperations group gets nervousbecause the entire plant could be shutdown if a second pump trips for anyreason. This means a significant loss ofproduction. Furthermore, the longrestart process generates a major flar-ing incident with its environmentalimpact. Thus, if one of these pumps,turbines or auxiliaries needs mainte-nance, it is considered an emergencyrequiring a 24-hour priority repaircrew.

A debottlenecking project wasimplemented to coincide with aplanned major maintenance outage.The project modifications necessitatedan increase in boiler feedwater flow.The obvious solution was to add afourth identical pump rated at 1375U.S.gpm and 2000 psig at 4310 rpm.Barrel pumps are expensive, however,and need costly foundations, highpressure piping modifications, steampiping modifications and a large instal-lation space.

To reduce costs, the actualflowrates were carefully examined.This current project required an addi-tional 300 U.S.gpm. Any foreseeabledebottlenecking project would requireonly an additional 750 U.S. gpm(including a contingency of 200 U.S.gpm). A fourth pump would have pro-vided too much flow. Thus, we investi-gated the possibility of rerating theexisting pumps to deliver 1750 U.S.gpm. Fortunately, the pump manufac-turer had a standard inner bundle thatcould provide the flow at 4390 rpmand would be interchangeable withthe existing barrel. The existing sealsand bearing could even be reused. Thesteam turbine manufacturer alsoadvised that the existing turbine couldbe rerated simply and inexpensively.

In all, the modifications were only 3/4the capital cost of new equipment.Once installation costs were consid-ered, the rerate option was the obviouschoice.

MODIFICATIONS PERFORMEDThe rerate resulted in the follow-

ing equipment modifications:

Pumps. The pumps are 6x8x10horizontal 9 stage API 610 barrel stylepumps with opposed impellers and aninternal crossover for thrust balance(Figure 1). The complete inner assem-bly consisting of the rotating elementand the inner volute case werereplaced. The outer barrel, bearingsand seals remained the same. Theseare the highest capacity internals thatwould fit into the barrel.

Turbines. The turbines were API611 single stage impulse type designthat used 500 lb inlet and 50 lb exhauststeam (Figure 2). The power rating wasincreased from 2100 to 2800 hp. Thisrequired a new nozzle ring and gover-nor valve. In addition, an electronicgovernor conversion, new-technologydry gas steam seals installation and ageneral overhaul were performed.These changes approached the maxi-mum power rating of the design.

Controls. The old control schemeconsisted of a hydraulic mechanicalgovernor that maintained a constantspeed of 4310 rpm and individual boil-er flow control valves. For energy opti-mization reasons, a new electronicgovernor control scheme was installedthat allowed the flow to vary by regu-lating the turbine speed from 3700 to4610 rpm while the header pressureremained constant.

Coupling. A new lightweightflexible disc coupling was installedwith a higher power rating.

Minimum Flow Valves. Newhigher capacity and updated modulat-ing valves were installed.

Although we had difficulties withall five items, this article discussesonly the major problems with the tur-bines and pumps in a chronologicalorder. For simplicity, the turbine issuesare grouped separately from the pump


HANDBOOK??Pump and Turbine Rerate Problems

Tests and observations point to geometry change as solution.By Leon K. Stanmore


problems, but you will need to try toimagine everything happening simul-taneously. Also keep in mind the diffi-culty in troubleshooting andperforming accurate analysis in anoperating plant environment. Com-pared to a test stand, we didn't haveenough instrumentation, we couldn'tcontrol all the variables, and we had toavoid disrupting plant operations.

EARLY OCTOBER: TURBINE STARTUPBoth turbines had been sent to the

manufacturers' service center toensure that the work was done proper-ly. They were installed and started upwithout any problems. Vibration lev-els were acceptable. But then, twounexplained trips on the B turbineoccurred. No cause was ever found forthese incidents. Vibration checks con-tinued to show normal readings.

EARLY NOVEMBER: B TURBINEPROBLEM 1

We then experienced four gover-nor valve failures over a weekend. OnFriday, the governor valve stem brokeinto three pieces. The guide bushingwas found inside the turbine casingand reused with a new stem from ourwarehouse. On Saturday the guide

bushing worked loose and machineditself through the seat. The valve stemwas worn and the governor valve pinhad sheared. A new valve seat andbushing were installed, but since thespare stem had already been used, theexisting stem had to be refurbished.The valve was installed on the stemwith a pin of unknown origin. On Sun-day the valve pin sheared, and thebushing was out of place. During thisfailure there was an upset with theelectronic governor of the other pumpresulting in a loss of 1500 lb steamheader pressure and a total plant shut-down. On Tuesday the valve stem wasso badly worn it was grabbing thebushing and limiting valve travel. Bythis time, all new OEM parts hadarrived on site, and a new seat, bush-ing, stem, valve and pin were installed.A mechanic noticed that the old bush-ing was magnetic and the new one wasnot, suggesting a material error. Theturbine was placed back into serviceand restarted without incident at nor-mal bearing vibration levels.

LATE NOVEMBER: B TURBINESOLUTION 1

The manufacturer was surprisedat first because governor stem failures

were extremely rare for this model tur-bine. But they immediately came upwith a solution. The original designhad a nitrided 416SS bushing and anitrided governor valve stem. In somerare cases, the differential expansionbetween the steel cage and the bush-ing resulted in a loss press fit. Theirsolution was a nitrided 304SS bushingthat had expansion rates higher thansteel and was self-locking. This mater-ial became the OEM standard in early1982. Since the new parts conformedto the new standard, the manufactureradvised that this would solve the prob-lem. The only nagging doubt was whythe original bushing had lasted for 18years. The overhaul records confirmedthat the bushings were checked andfound to be in excellent condition. Theexplanation offered was that the rerat-ing resulted in increased velocityacross the valve. This resulted in high-er vibration that loosened the bushing.

LATE DECEMBER: B TURBINEPROBLEM 2

Just before Christmas an operatorpassing by heard an unusual squealemanating from the turbine. By thetime we arrived with our vibrationinstruments, everything was quiet andvibrations were normal. We decided todo some testing. The B pump was puton manual and the speed was variedthrough the operating range. The Apump was left on automatic and itsspeed adjusted to provide the requiredflowrates. A dual channel analyzerwas programmed for automatic datacapture every 100 rpm. A resonancewas discovered around 4000 rpm andreproduced the squeal. The vibrationenergy fluctuated but was extremelyhigh, and occurred with a specificspike around 162,000 cpm (Figure 3).It was measurable all over the turbinebut was highest around the governorvalve area (in the 3 in/sec and 10grange). When the turbine was operat-ing at resonance speed, one could seevibrations causing the trip lever towalk off the knife edges. We conclud-ed that this resonance was the rootcause of the mysterious trips and thebroken governor valve stems. Theknife edges were not in particularlygood condition. They were replacedand the turbine placed on manual at4200rpm, where vibration levels werenormal and more analysis began. Sim-ilar vibration resonance patterns werenoticed on the A pump but with muchsmaller vibration amplitudes.

Figure 1. Pump cross section

Figure 2. Turbine cross section


MID JANUARY: B TURBINESOLUTION 2

Our analysis pointed to problemswith the nozzle ring/governor valvecombination. The manufacturers'rerate calculations consist of a simpletable showing maximum power, plus amargin, versus flow-passing capabili-ties of various nozzle ring andvalve/seat combinations. Our horse-power requirements nudged us intothe largest valve/nozzle ring combina-tion available for this turbine. Howev-er, further investigations revealed thatthis table was based on standardmechanical governor data with a max-imum valve travel of 0.6". The newelectronic governor used a controlvalve actuator with a permissible trav-el of 1". We had awarded the governorconversion to the turbine manufactur-er specifically to avoid this sort of con-fusion, but it happened anyway.

The new two-stage governor valvehad an aggressive 40º front taper anda 45º rear angle. It could actually pro-duce 3052 hp at a stroke of 0.43". Themanufacturer had a valve with a gen-tle 20-30º valve angle that could pro-vide the power with 0.675" stroke.This valve had proved to be "more sta-ble" in other applications, but deliverywas six weeks. Interestingly, our origi-nal 30-30º valve could meet the powerrequirements with a 0.625" stroke.The theory was that a different valvewould operate at a different position,one that would change the velocity dis-tribution and the exciting force andwould thus move the resonance pointaway from our operating range. Unfor-tunately, the manufacturer could notcalculate these resonance relation-ships, so we could only postulate anduse the trial-and-error method.

LATE JANUARY: B TURBINEPROBLEM 3

We felt that the valve position the-ory was sound, and since the old 30-30º valve was still in our warehouse,we decided to reinstall it and reevalu-ate. We were extremely disappointedwith the results. Although the vibra-tion levels were marginally lower, theresonance persisted at the same fre-quency and now peaked at 4100 rpm.

EARLY MARCH: B TURBINESOLUTION 3

We performed another test, mea-

suring inlet and outlet pressures andtemperatures, steam chest pressures,valve stroke, vibration levels andspeed in an attempt to correlate stroketo critical pressures. This was unsuc-cessful. It appeared that we had to putthe turbine on manual away from theresonance and forgo the energy savingbenefits and the major justification forthe electronic governor conversion.But having come this far, we decidedto try one more time and install themore stable 20-30º valve, which hadfinally arrived on site. Note that the 20-30º valve is identical to the 40-45ºvalve, except for the machining of thevalve angle. The 30-30º valve is a dif-ferent design and casting (Photo 1).

Although our enthusiasm wasdampened, we installed the 20-30ºvalve and made a complete testthroughout the speed range. We werepleasantly surprised to find that alltraces of the problem disappeared.Vibration levels were 0.1 in/sec withno signs of resonance. We were able toturn over the unit to Operations with-out restrictions.

LATE JUNE: A TURBINEPROBLEMS

As you will see in Part 2 of thisarticle, we were also having problemswith the pumps. Internal leakage hadcaused the speed to increase steadilyover the months until the equipmentwas operating at 4500 rpm. The pumpproblems also produced higher vibra-tion levels, and these suddenly beganto increase exponentially. In anattempt to minimize damage until therepair parts arrived, we asked Opera-tions to reduce the discharge headerpressure setpoint to the absolute mini-mum. This was successful, and overallvibration levels did decrease.

Shortly thereafter, the B pumpwas repaired and the operating speeds

of both pumps dropped dramaticallyto 4005 rpm. This just happened tocoincide with the turbine governorvalve resonance speed that we had dis-covered earlier and repaired on the Bturbine. We were planning to replacethis valve on the A turbine with thenew 20-30º design during the pumpoutage. Overall vibration levels nearthe steam chest had been increasingbut had doubled in a week to the 20 grange. We suspect that over themonths, the bushing had worn andwas now rapidly deteriorating. In anattempt to move the turbine out of theresonance range and limp along untilthe pump parts arrived, we askedOperations to raise the header pres-sure to increase the speed out of theresonance range. Turbine vibrationlevels did diminish, but several opera-tors expressed bewilderment aboutwhat we were doing.

LATE JULY: A TURBINESTARTUP

A new governor valve, stem, pin,seat and bushing were installed on theA Turbine. It was now identical to therepaired B unit. The turbines startedsmoothly with overall vibration levelsbelow 0.1 in/sec., and there were nosigns of resonance.

PART 2You may recall that Part 1 dealt

mainly with turbine issues as part ofthe Novacor Chemicals boiler feed-water rerate project. For simplicity,Part 2 discusses primarily pumpissues. In reality, however, the chal-lenge was considerably more com-plicated because pump and turbineproblems were interrelated.

EARLY OCTOBER: PUMP STARTUPAs soon as the pumps were

started, we knew we had problems.

Figure 3. Waterfall spectrum of OB turbine bearing before modifications (vertical direction)


0.8

0.4

0.0

In/s

pc

3600

4000

1.5K 150K 300K cpm

4400r pm

On minimum flow bypass, overallvibration levels exceeded 0.8 in/secin the vertical direction on theinboard bearing housing of bothpumps. And, even after more nor-mal flows were established, overallvibration readings were 0.4 to 0.6in/sec. When the situation was ana-lyzed, it became clear that all theenergy was coming from peaks thatwere at 5x and 7x running speed(Figure 4). Note that the new pumpinternals had a seven-vane impelleron stages 5 and 6 and five-vaneimpellers on the other stages. The Bpump had higher vibration levelswhile the A pump also exhibitedlower energy peaks at 3x, 4x and 6xrunning speed, indicating some kindof looseness (Figure 5). There waslittle vibration on the outboardbearing. This is perhaps explainedby the fact that the throttle bushingstabilized that end of the pump.

We approached the manufactur-er for assistance. Its representatives'

first reaction was that this was not aproblem and we should not be con-cerned. They provided some litera-ture advising that operation with highfrequency blade pass vibrations didnot pose a problem for long term reli-ability. Our position was that a newpump should meet the specified API610 vibration levels for acceptance.

LATE DECEMBER:WORSENING PUMP PROBLEMS

Vibration data was continuous-ly supplied to the manufacturer forreview and comment. Overall vibra-tion levels steadily increased whilethe amplitude of the 5x and 7xpeaks appeared to change over time.We reported operating concerns totheir engineering group all alongbut felt it was important to formal-ize our position. We therefore filed awarranty claim. This elevated theproblem to their senior manage-ment, and we began receiving moreactive attention.

LATE JANUARY:PUMP INBOARD BEARINGRESONANCE TESTING

We thought that the evidenceclearly showed a blade pass prob-lem. But the pump manufacturersuspected a resonance problem inthe bearing support. Apparently,problems of this nature had beenreported previously. They agreedthat vibrations were caused by vanepass pressure pulsations, but the realproblem was that these pulsationswere being amplified by a bearinghousing/bracket resonance. Theirsolution was to increase the naturalfrequency of the inboard bearinghousing. We agreed to do some test-ing to validate this theory. Unfortu-nately, we did not have a calibratedhammer, so the results are not highlyaccurate. They did indicate the pres-ence of a resonance, however.

The B pump was shut down, anda vibration probe was placed on theinboard bearing housing in the verti-cal plane. A baseline signature wasrecorded. Then the housing wasstruck with a rubber mallet numer-ous times. The frequency analysisshowed a definite spike occurring at25,200 cpm (Figure 6). The processwas repeated with the vibrationprobe in the horizontal direction. Aspike was also noticed at 23,175 cpmbut at half the amplitude. This corre-lated with the field vibration mea-surements that always showed

Figure 4. Typical frequency spectrum of B pump before modifications

0K 30K 60K cpm

VERTICAL

HORIZONTAL

5X 7X0.6

0.3

0.0

0.2

0.1

0.0

In/s

pc

In/

s pc

Figure 5. Typical frequency spectrum of A pump before modifications

Figure 6. Results of resonance testing ofinboard pump bearing housing

150 30K 60K cpm

0K 30K 60K cpm

0K 30K 60K cpm

VERTICAL

HORIZONTAL

A – BASELINE SPECTRUM

B – NUMEROUS IMPACTS WITHRUBBER MALLET – 32 AVGS

1X 2X 3X 4X 5X 6X0.8

0.4

0.0

0.4

0.2

0.0

0.4

0.2

0.0

0.4

0.2

0.0

In/s

pc

I

n/s

pc

In/s

pc

In/s

pc


higher readings in the vertical direc-tion. For our operating speed range of3700-4500 rpm, the 5x range is 18,500to 22,500 cpm, and the 7x range is25,900 to 31,500 cpm. This couldexplain the strange variance of the 5xand 7x spikes in our field testing.

Weights were added to the bear-ing housing in an attempt to alter itsnatural frequency. A long threadedrod was installed in the bearing hous-ing lifting eyebolt tap, and first 65 lbsand then 100 lbs of lead weights werebolted in place. The pumps were runthrough the operating speed rangeand vibration levels were comparedto the baseline. We discovered that100 lbs significantly changed the over-all vibration levels, and the magni-tude of the 5x and 7x frequencyspikes as the speed varied (Figure 7).

EARLY MARCH:PUMP REPAIR RECOMMENDATIONS

The manufacturer reps felt thatonly bearing bracket modifications

were needed to reduce vibrations toacceptable levels. Their calculationsshowed that if the mass of the bearinghousing were increased by 400 lbs, itwould move the natural frequencycompletely outside the operatingrange. But we felt this was an exces-sive amount of mass to add physical-ly. Therefore, it was decided toincrease the stiffness of the bearingbracket. Their calculations showedthat a 150% increase would alsosolve the problem. Their proposalwas to disassemble the pump,remove the inboard end cover andcut off the welded bearing housingsupport. They would supply a newsupport that was 250% stiffer thanthe old one, weld it onto the old endcover, stress relieve and reinstall.The total turnaround time to disas-semble, send the parts to their repairfacility and reassemble was 2-3weeks. If these bearing bracket stiff-ening modifications did not work,they would then pursue hydraulic

modifications. We had a problem with the pro-

posal. First, an outage of this lengthwas unacceptable to our OperationsDepartment. Second, plant opera-tions had already been severely dis-rupted by all the trips, shutdownsand tests that were performed. Theentire pump would have to be disas-sembled to gain access to the inboardend cover. The costs associated withpump disassembly are very high,and we did not have the luxury of alengthy test program. We needed abetter solution.

Resonance test results indicatedthat bearing housing modificationswere necessary. But we felt that weshould also address the hydraulicproblem. The first issue was impellerstack-up. Multistage pumps withidentical trailing edge vane positionsare known to amplify blade passvibration energy. Although norecords were kept, we were assuredthat the impeller keys are carefullypositioned to avoid vane line-up, andthat this item is always checked aspart of their normal quality inspec-tion process.

Another problem area was theimpeller blade tip-to-volute tonguegap. API 610 recognizes that highFigure 7. Waterfall spectrum showing effects of adding weights to pump inboard bearing

housing

A – NO WEIGHTS

B – 100 LB WEIGHT ADDED

0.6

0.3

0.00K 30K 60K cpm

3700

3700

4100

4500

4100

r pm

rpm

0K 30K 60K cpm

0.4

0.2

0.0

In/s

pc

In

/s p

c

Figure 8a. Impeller/volute trim detail(stage impellers)

Figure 8b. Impeller/volute trim detail(1st stage impeller)

a = 1/2 in.b = 1/4 in.

stageimpellers

1st stageimpellers

volutetongue

volutetongue

ba


energy pumps require special provi-sions to reduce vane passing fre-quency vibrations, and the standardspecifies that the gap be at least 6%.Our as-built gap of 6.8% was margin-ally acceptable. We asked the manu-facturer to develop a repairrecommendation,and they increasedthis gap very creatively. Instead oftrimming the impellers and underfil-ing to restore the head, they trimmedthe impellers at an angle. The anglecuts also were in the opposite direc-tion of the volute trims. The theorywas to create a "scissoring" actionthat would not only increase theeffective gap but change the energydistribution over the volute tongue tolessen the magnitude of the pulsa-tions. The eye sides of the impellertips were cut back about 1/4" and thevolute tongues 1/2" in the oppositedirection. The first stage was a dou-ble suction, and application of thattheory resulted in "V" cuts (Figures8a and 8b).

We decided to pursue hydraulicmodifications and bearing bracketstiffening simultaneously. To mini-mize repair time, we purchased anew end cover. Unfortunately, it wasan expensive forging that had a four-month delivery cycle. The pumpmanufacturer agreed to pay for thebearing bracket modifications. Weagreed to pay for a new end coverand the hydraulic modifications. Allmodification work would be done atthe pump manufacturer's local ser-vice center.

We had purchased a completeinner bundle as a spare to upgradethe electric pump at a future date.This unused bundle was the first tobe modified. When it was opened,we discovered that the impossiblehad occurred. The vanes of the two7-vane impellers were exactlyaligned. Four of the 5-vane impellerswere exactly aligned. The three oth-er 5-vane impellers also were alignedin a different plane. This situationrequired that all impellers beremoved and their keyways weldedup and recut. Also, two impellersseized on the shaft during disassem-bly, and it had to be undercut andchrome plated.

LATE MAY:PUMP OPERATING PROBLEMS

We had planned to do the A

pump first because it showed signs ofmechanical looseness. However, wenoticed a deterioration in pump per-formance. The pumps were slowlyspeeding up. At first we thought theprocess designers had made a mis-take, and we required more flowthan planned. However, in Decem-ber we began noticing a strange butmoderate cavitation noise in the bal-ance line near the first elbow. Overmonths of negotiations and testing,this noise became louder.

We started to trend vibrationlevels on the balance line near thethrottle bushing and noted a steadyincrease. By the end of May, thevibration energy was 10 g. In earlyJune it was 20 g. When all the partsfinally arrived in the first week ofJune, it was more than 30 g. Bythen, we had decided to repair the Bpump first and expected to see someinternal damage. We were not disap-pointed.

EARLY JUNE:B PUMP REPAIR

Severe cavitation damage wasdiscovered when the pump's out-board end cover was removed. Thethrottle bushing and sleeve are over-laid with welded stellite material. Thesleeve on the shaft had contacted thestationary bushing. They had weldedthemselves together and spun in thesupport. This opened a leak path, and

the interstage pressure of 1000 psihad cavitated the 3/4" housing almostto failure (Photo 1). During startupwe experiences some system upsetsthat might have caused excessiverotor movement and contact. Anoth-er possible explanation was improp-er installation. We discovered thatthe bearing housings had not beenrealigned after the new internalswere fitted. To reduce costs, the man-ufacturer's service representativeswere not employed during the origi-nal installation. But they were on sitefor all rebuilds.

The end cover was shipped tothe manufacturer's repair center tohave the washed out holder cut offand a new ring welded on. A newthrottle bushing was installed, andthe repaired end cover was back onsite within 3 days. The new suctionend cover with the heavier bearingsupports was installed (Photo 2), andthe bearing housings were carefullyaligned. The new pump bundle wasfitted without incident.

MID JUNE:B PUMP STARTUP

We had found mechanical dam-age, corrected the impeller stack-up,made hydraulic modifications andstiffened the support. The pumps hadto work because there was nothingleft to be done. But quite honestly, wewere nervous because of the history

Figure 9. Waterfall spectrum of B pump inboard bearing after modifications

rpm

In/s

pc

15K 25K 35K cpm

0.4

0.2

0.0

3700

7X

5X

4100

4500


of events. The B pump was run up tominimum governor of 3700 rpm withthe electric pump still on and the Apump on automatic. The B pump wasmanually run up in 100rpm incre-ments ,and vibration spectrums weretaken. Then the electric pump wasshut down and the speed lowered in100 rpm increments, and vibrationreadings were taken.

Success was nearly complete.Overall vibration was down to accept-able levels. However, spectrum analy-sis showed that the vibration energywas still coming from 5x and 7xpeaks, albeit at much lower levels. Asspeeds were lowered and the pumpwas pushed back on the curve, the 5xspike dropped off, but the 7x peakincreased to 0.4 in/sec at 3900 rpm,indicating that there was still somesort of resonance, Figure 9. Also, therepaired B pump now operated at 85rpm faster than the A pump, indicat-ing that the underfiling did not restorethe head expected. Results were shortof perfect, but the unit was turnedover to Operations without restric-tions. Shortly after, the pump was shutdown for a minor turbine repair, andthe bearing housing resonance testwas repeated. The previous verticalresonance had all but disappeared.

MID JULY:A PUMP REPAIRS

The bundle removed from the Bpump was sent to the service centerfor inspection and hydraulic modifi-cations before installation into the Apump. With all the cavitation to theend cover, we thought there wouldbe major internal pump damage aswell. Surprisingly, there was littlerotor degradation except for theseized throttle bushing. There wasslight wear on the interstage bush-ing. Both were replaced. The otherwear rings were in excellent condi-tion. The impeller wear rings werespecified with API 610 clearances of0.018 in., but the pump manufactur-er insisted that the throttle and inter-stage bushing clearances remain attheir standard of 0.012 in. Again, wefound an impeller stacking problem:four 5-vane impellers lined up. Twoimpellers also seized on the shaftduring disassembly. A temperatureprobe, accidentally installed in theunused vibration probe location, cuta groove into the burnished surface.

All these difficulties were correctedfor, and a heavier bearing bracketwas installed on the old end cover.

The pump was then shut downand disassembled. We searched forsome looseness that could explainthis pump's abnormal vibration sig-nature. We were almost relieved tofind that a loose journal bearing fit inthe top inboard housing had causedsevere damage to the babbited sur-face. The measured clearance wasmore than 0.004in. At speeds exceed-ing 4000rpm, the manufactureradvised that their normal practice isto shim this clearance to 0.0005 in.Either the fit had worn or the shimswere accidentally left out duringreassembly.

LATE JULY:A PUMP STARTUP

The rest of the pump rebuildand turbine governor valve, seat andbushing installation went withoutincident. The pump bearing hous-ings were carefully realigned. Ourcrew was very efficient by this time.Startup was enlivened by a pluggedoil passage in the pump inboardhousing, a siutation that wiped thejournal bearing.

We were pleased to report satis-factory results. But again the start-up tests showed signs of aresonance at 3900 rpm with 5x and7x spikes approaching 0.3 in/sec.And at speeds below 4000 rpm,small but noticeable 6x and 8xpeaks started to appear. At higherspeeds everything smoothed out,and overall levels were below 0.2in/sec. The pump manufacturer sus-pects some sort of acoustical reso-nance is taking place. Again,everything was not perfect, but theunit was turned over to Operationswithout restrictions.

The bundle that was removedfrom the A pump was sent to themanufacturer's service center forinspection and hydraulic modifica-tions before being returned as awarehouse spare. Again, there wereimpeller vane line-up problems.There was also some erosion wearon one side of the throttle and inter-stage bushings. Obviously, we hadsimilar problems to the cavitatedpump and were just lucky thatthere was no physical contact. Bothitems were replaced.

CONCLUSIONSThis rerate was presumed to be

a straightforward replacement ofproven parts with complete assur-ances from original equipmentmanufacturers. What went wrongand what can be done to ensure itdoes not happen again? Unfortu-nately, there is no simple answer tothis question other than the factthat rerates of complex turboma-chinery can be difficult, particular-ly when resonances are involved.These can be extremely difficult topredict, isolate and correct. To solvethese problems, extensive effortand cooperation between our tech-nical staff and the manufacturerswas necessary. Also, we needed agreat deal of patience and under-standing from our Operations andMaintenance departments. As aresult of our efforts, we did manageto solve some complex technicalproblems together, and the pumpsand turbines are now operating sat-isfactorily.

Following are some specific rec-ommendations resulting from thisproject:

• This experience reinforces theneed for shop testing wherever pos-sible. It can eliminate many fieldproblems and allow more accurateanalysis and easier modifications.We actually tried to procure a testbarrel for the pumps, but it provedto be prohibitively expensive. Weshould have insisted that the tur-bine be sent to the nearest OEMfacility with a test loop and paid thepremium.

• We will change our multistagepump specifications to require aninspection hold point for impellerstack-up verification. We had hireda third party inspector and used theOEM's superb quality plan and stillmissed this important item.

• For all critical equipment, wewill now specify an engineeringaudit of manufacturer's rerate cal-culations and selections. We used todo this only for our large compres-sors.

• For major rerate work, use ofthe OEM service representative isrecommended as inexpensive insur-ance.


• We need a more open rela-tionship with our OEM suppliers.This is especially true for reratework , in which reliability is moreimportant than capital costs. Wethought that we had unique prob-lems but discovered afterwards thatother users had faced similar issues.With a more open dialogue beforethe order was finalized, we couldhave minimized the adverseimpact.

ACKNOWLEDGMENTSI would like to recognize and

thank all those who participated inthis project. Special thanks go toJim Gardiner, a rotating equipmentengineer who has since left Nova-cor and is now working for NorthAtlantic Refining Co., and to MikeDufresne and Fred Robinett ofSulzer Bingham and to Rob Kirk-patrick from Elliott Turbomachin-ery Canada. ■

Editor's Note:This article is reproduced with

permission of the TurbomachineryLaboratory, from Proceedings of the13th International Pump Users Sym-posium, Turbomachinery Laboratory,Texas A&M University, College Sta-tion, TX, pp.89-95, Copyright 1996.

John P. Henderson is an associateengineer with Novacor Chemicals'ethylene plant in Sarnia, Ontario. Heis a rotating equipment specialistresponsible for technical support,troubleshooting, reliability improve-ments, vibration analysis and newequipment review and specification.He received a B.S. degree in mechani-cal engineering from Carleton Univer-sity in Ottawa and is a RegisteredProfessional Engineer in Alberta.


Throughout our lives we facehidden dangers. When we wereyoung, it was the monster under ourbed. In our late teens it was a patrolcar in the rear view mirror. Today, wework on pumps with more hiddendangers than a mine field. One of thebiggest is cracking in various pumpcomponents: shafts, impellers andcouplings on the rotating end; cases,nozzles and fittings on stationaryparts. When do you look? How do youlook? And who should look for thesepotentially dangerous situations.

Maintenance staffs are expectedto do more with less in this era of"right sizing."

Every step of our process must beproved cost effective and essential foroperation. Some maintenance stepsare being dropped entirely, and othersare being cut back. So, with fewerfinancial resources available and ashorter time period to work on equip-ment, where to we concentrate?

Some problems, such as wornor broken parts, bearings and seals –are obvious, of course. But otherdifficulties are harder to detect. Itis impossible to see a difference of.002" at a bearing fit but it coulddetermine whether the pump keepsrunning or not. We cannot tell byvisual inspection a difference of.005" or .006" over specs in animpeller wear ring clearance, butit could affect pump performance.So why do we visually inspectpump parts? It is easy to say every-thing looks okay, change the bear-ings and seals and put the pumpback together. When you do this,what risks are you taking? If it is asmall ANSI pump, there isn't muchof a risk. Damage can result, but thecost would be small. On the otherhand, if we are dealing with a largeboiler feed water pump, damagefrom a hidden crack could result in

a major repair expense and extend-ed down time.

In the utility industry, downtime is money. In the petrochemicalindustry a failed pump may result infar more than simply spilling somewater on the floor. Fire, explosionand death could occur.

Figure 1 shows an inspectionreport on a large boiler feed pump.Under the results, item #1 is the suc-tion impeller. If this impeller hadcome apart, it would have propelledmetal fragments through the pump.Because the pump runs at 5600 rpm,the damage and mess would be con-siderable. The cracks on this particu-lar pump were visible to the nakedeye and easy to spot. Item #5 is a dif-ferent story, however. Without thewet particle inspection and the eddy-current test we would never knowthese cracks existed.

I'm not going into what causescracking. This subject would beworth an article itself. What I wantto stress is the importance of havinga good inspection system in place.

There is an old saying..."Pay me now,or pay me later." This fits the need toinspect to a tee. ■

Gary Glidden is crew leader of thepump shop at Houston Lighting andPower and a member of the Pumpsand Systems User Advisory Team.

Looking for Hidden DangersBy Gary Glidden


HANDBOOK??

Example of a boiler feed pump's impeller damage that started as a minute crack.

PHOT

O CO

URTE

SY O

F HO

USTO

N LI

GHTI

NG A

ND P

OWER


Figure 1. Sample inspection report



HANDBOOK?

IntroductionPatient: “Doctor, it hurts when I dothat.”Doctor: “Well, don’t do that.”

This old vaudeville line can stillmake a point with regard tolow Net Positive Suction Head(NPSH) conditions in a pump-

ing system. The very existence ofsuch a condition is forgivable only inthe few situations where it truly can-not be avoided at the system designstage. Low NPSH situations are occa-sionally a necessary evil in processconditions, or in a batch operation inwhich the function of the pump is toempty a tank. In the latter situation,the pump is bound to cavitatetowards the end of the batch, wherethe NPSH steadily decreases belowthe value required by the impellerfor satisfactory operation.

In considering the variety ofways out of a low NPSH, or cavitat-ing condition, we should first under-stand the source of the problem –but without getting into an exhaus-tive treatise on the subject.

Pumps cavitate when the NPSHrequired by the impeller is greaterthan the NPSH being made availablefrom the system in which it operates.Under these conditions the liquidvaporizes in the eye of the impeller,and the bubbles created then col-lapse in a series of implosions on thevanes.

In trouble-shooting cavitation

problems, it is important to recognizethat the physical symptoms thatidentify cavitation difficulties alsoreveal other conditions which aretotally unrelated. But these will bediscussed in a future article.

To stop or prevent cavitation, wehave two options:

1. decrease the NPSH required by the pump

2. increase the NPSH available from the system

The PumpThe NPSH required by the

pump is a function of the hydraulicdesign of the eye area of the impeller.The level of NPSH required can bereduced by increasing the eye area ofthat impeller. However, this can ren-der the impeller more susceptible tosuction recirculation, with almostidentical symptoms as cavitation.Consequently, I would recommendthat the pump manufacturer be con-

sulted in any change of this nature.Some manufacturers will sometimeshave an alternative impeller, with alower NPSH requirement, availablefor that particular pump.

Another possibility that hasbeen used with varying success isthe suction inducer. It consists of asmall axial flow screw arrangementin the eye of the impeller to give theliquid a pre-rotation. This raises thepressure just enough to preventvaporization from taking place in theeye of the impeller. The inducer isnot always available as very fewpump manufacturers offer thisoption.

The only other choice in reduc-

The Correction andPrevention of LowNPSH Conditions

Understanding the source of the problem will identify the way out.

By Ross C. Mackay

NPSHAvailable

fromSystem

NPSHRequiredby Pump

To Cavitation

Figure 1. NPSH balance diagram

Photo 1. Double suction impeller pump

(COU

RTES

YOF

PATT

ERSO

NPU

MP

CO.)


ing the NPSH required is a completechange of pump style or size. Hereare some options:

• A lower speed pump requiresless NPSH but will need a largerimpeller to handle the same pump-ing conditions.

• As a result of the two eyeareas, a double suction impellerdesign (Photo 1)needs only two-thirds of the NPSH required by asimilarly sized single suction design.

• A vertical canned pump (Fig-ure 2) can have additional static suc-tion head built into the columnlength to allow for the needed NPSH.

• A number of lower flowpumps operating in parallel will alsoneed less NPSH.

Another solution involves theuse of a booster pump upstream ofthe main pump. The booster unitmust operate over the same capacityrange, but it can develop a lowerhead. In this arrangement it needs to

develop only the amount of headthat is needed for the NPSH requiredby the main pump. It can, therefore,be a low head and/or low speedpump, both of which need lessNPSH.

The reader will notice that of theseven options affecting the pumpside of the equation, one option cancreate a similar adverse condition,one option is not widely available,and the other five options requirenew pumps. Perhaps we should con-sider the system side of the equation.

The SystemThe NPSH Available from a sys-

tem consists of the following fourfactors only. Consequently, at leastone of these must be changed toincrease the NPSH Available (Figure3).

NPSHA = Hs + Ha - Hvp - Hf

the static head over the impellercenterline (Hs)

the pressure on the surface of the liquid (Ha)

the vapor pressure of the liquid (Hvp)

the friction losses in the suction line (Hf)

It is apparent that the effectivecures are those that increase the firsttwo factors in the equation ordecrease the last two.

Static HeadTo increase the positive static

head is a simple (?) matter of lower-ing the pump or raising the suctiontank, or raising the level of the liquidinside the suction tank. While thephysical movement of the tank or

pump would often be an expensiveproposition, the raising of the tanklevels may be relatively cheap andsimple. However, lowering the pumpcan often be more economical whentaken together with any otherchanges that are being effected ifmore than one problem is being cor-rected. For example, if the suctionpiping arrangement is being changedto stop the creation of turbulence inthe inlet, the pump could be movedto a lower floor of the building. Thiswould straighten the inlet piping,eliminate the turbulent flow andincrease the static head.

Surface PressureSurface pressure is a little tricky

to change if the suction source is theAtlantic Ocean or some other bodyof water that resists being controlledby mere mortals. It might be possi-ble, however, to enclose a man-madetank and pressurize it, or even intro-duce a nitrogen blanket. Both ofthese possible solutions are subjectto the dictates of the operating sys-tem. For example, increasing thepressure inside a deaerator woulddefeat the whole function of that ves-sel and thus must be judged imprac-tical. But since this pressure is one ofonly four factors in the NPSH formu-la, it is worthy of some considerationin certain installations.

Vapor PressureThe only way to reduce the

vapor pressure of a liquid is toreduce its temperature. In manyinstances this is operationally unac-ceptable and can be ignored. Also,the extent of the temperature changeneeded to provide an appreciable dif-ference in NPSH usually renders thismethod inappropriate.

Friction LossesOwing to the fact that pump

inlet piping is notoriously bad in thevast majority of industries through-out the world, this is the area wheresignificant improvements can oftenbe realized. However, I must cautionyou against the tendency to shortenthe length of suction piping simply toreduce friction losses. While this willbe effective, it could deny the liquidthe opportunity of a smooth flow

Figure 2. Vertical canned pump

Hs

Hf

Hvp

Ha

Figure 3. NPSH diagram


DRIVER

DISCHARGEHEAD ASSEMBLY

SUCTION

BARREL

BOWLASSEMBLY

PACKINGBOX

DISCHARGE

OIL FIELD PRESSURIZATION PUMPHIGH PRESSURE

MULTIPLE PUMPS CAN BE USED IN SERIESBARREL OR CAN PUMP

path to the impeller eye. This, inturn, could cause turbulence andresult in air entrainment difficultiesthat create the same symptoms ascavitation. To avoid this, the pumpshould be provided with a straightrun of suction line in a length equiv-alent to 5 -10 times the diameter ofthe pipe. The smaller multipliershould be used on the larger pipediameters and vice versa.

The most effective way of reduc-ing the friction losses on the suctionside is to increase the size of the line.This can make a dramatic differencein the losses, to the extent that morethan a 50% reduction in friction canbe realized by replacing a 12 inchline by a 14 inch line. Exchanging a 6inch line for an 8 inch line canreduce the friction losses by as muchas 75%. It must be acknowledged,however, that changing the pipe sizealso changes the size of all valves andfittings.

Reductions in friction losses canbe achieved even with the same linesize by incorporating long sweepelbows, changing the valve type andreducing their number.

One final item that shows upwith surprising regularity is the suc-tion strainer. It is widely used in the

commissioning stages of a new plant.Unfortunately, there are many timeswhen it is a forgotten piece of equip-ment, and blockage in the strainerbasket gradually increases and raisesthe friction loss to an unacceptablelevel. If a strainer is required in asystem on a continuing basis, itshould be located downstream of thepump, and the pump should be onethat can handle the solid sizesexpected. Obviously there are limita-tions to this concept, but it can beused more frequently than currentpractice would have us believe.

Most EffectiveThe most effective cure for cavi-

tation is the one which is both eco-nomically and practically viable. It isimportant to consider all possibilitiesand not dismiss any of them out ofhand just because they are going tocost a “lot of money.” If the cost ofrepair is, say, $25,000, and the pre-sent cost of nursing the pumpthrough its cavitation problems is$5,000 annually, it follows that itwould take five years to retrieve thecost of the repair. However, if youare planning to run that pump for anadditional 10 years after that date,the savings will amount to $50,000

by the end of the 15-year period. Bythe same token, if the cost of therepair is only $2,000, the decisionbecomes much simpler as the pay-back is achieved well within the firstyear.

From my experience in plants allover the United States and in manyother countries around the world, ithas become evident that manypumps are cavitating for the want ofone or two feet of NPSH. As a resultof this, I have found that the relative-ly simple change of raising the levelof the liquid at the suction source tobe an extremely effective and eco-nomical cure. Again, it is a little diffi-cult if the suction source happens tobe the Atlantic Ocean.

The best cure of all is, of course,prevention.

If the patient doesn’t ‘do that’ inthe first place, it will never hurt!■

Ross Mackay is a consultant inpump reliability. He has more than 35years of experience in the internationalpump community and conducts theMackay Pump School throughoutNorth America.



HANDBOOK??

Amunicipality purchased sever-al units to pump water on anemergency basis for fire fight-ing. These units are located

around the downtown area and canprovide water through portable pipesand hoses in case a major catastrophecuts off the primary water supply tofire hydrants. Each unit consists of aV-12 diesel engine driving a five stagevertical centrifugal pump through aright angle gearbox with a 1:1 ratio.The 1600 hp engine has an operatingspeed range of 1200 to 1800 rpm. Theengine flywheel connects to the gear-box through a drive shaft that has uni-versal joints on each end.

This article describes how failuresof the water pumping system werelinked to torsional vibration. It alsoshows how a solution was developedusing computer analyses. Field testswere performed to verify that thesolution was adequate to prevent addi-tional failures.

The ProblemThe system experienced failures

of the input gear and cooling fan. Sev-eral of the bolts that held the gear tothe input shaft broke after only 13hours of operation. It was speculatedthat this failure was due to excessivetorsional vibration or improper fit thatmay have caused the transmittedtorque to be unevenly distributedamong the bolts. The cooling fan ontop of the gearbox also experience sev-eral failures. The first time the fanblades broke, it was thought to berelated to possible problems with thematerial or manufacturing process.The fan is constructed by pressing the

general shape out of the sheetmetal, and then the blades aretwisted to the proper pitchangle. If a small crack formedat the base of the fan blades, ahigh stress riser would be cre-ated. However, this type offan has been used successful-ly at other installations. Thesefan blade failures could havebeen caused by high torsionalvibration.

Although it was possibleto use additional bolts to holdthe gear more securely to theinput shaft, there was con-cern that if torsional vibrationlevels were too severe, thensome other portion of the systemwould fail resulting in more damage.Therefore, a detailed torsional analy-sis was performed. The wet pumpimpeller inertial (25% greater thandry) was used to account for the water.The dynamic torque produced by theengine was calculated from the cylin-der pressures and inertia forces actingon each of the six throws. Appropriatestress concentration factors wereapplied to the gear and pump shaftsections. The damping in the systemwas evaluated to include the engineviscous damper properties as well asthe effects of the bearings, packingand pumped fluid.

The torsional analysis indicatedthat the engine produces significantdynamic torque at 2.5x, 3.5x, 4.5x and6x running speed. The interferencediagram in Figure 1 show the calculat-ed torsional natural frequencies andharmonic speed lines associated withthe engine. The intersection points

indicate torsional critical speeds. Thetwo circled points in Figure 1 showwhere the 4.5x and 3.5x engine har-monics intersect the fifth torsionalmode - at approximately 1400 and1800 rpm respectively. The fifth tor-sional mode was of concern since theshape showed twisting in the inputgear shaft and oscillation at the cool-ing fan. Also, the engine damper is notas effective for this mode. Steady-stateforced calculations were performed topredict the level of torsional vibrationversus engine speed. The alternatingshear shaft were above the endurancelimit of the shaft material. The forceanalysis also showed that the coolingfan would experience high torsionaloscillation.

Torsional Analysis YieldsSolution

Based on the torsional analysisresults, the system needed to be mod-ified to reduce the excessive torsional

Torsional Vibration Linked toWater Pumping System Failure

By Troy Feese


Figure 1. Interference diagram

vibration and stress levels. A paramet-ric study was performed. This showedthat using a “torsionally soft” couplingbetween the engine and gearboxwould help isolate the engine harmon-ics from the rest of the system. A cou-pling with a rubber element in shearwas selected that would have a lowtorsional stiffness (less than 300,000in-lb/rad) and provide additionaldamping to the system. Since this cou-pling bolts directly to the engine fly-wheel and supports the drive shaft,the system could be easily retrofittedin the field. The analysis indicatedthat with the rubber coupling the tor-sional vibration and stress levels in thegear and pump would be significantlyreduced compared to the original sys-tem.

Field measurements of the sys-tem were performed before and afterthe rubber coupling was installed toverify that the torsional vibrationwould be reduced enough to preventadditional failures. Strain gage mea-surements could not be made on theinput gear shaft due to limited accessi-bility inside the gearbox. However, thetorsional analysis showed a correla-tion between the predicted dynamictorque in the drive shaft and the alter-nating shear stress in the input gearshaft. Therefore, strain gages wereattached to a uniform section of thedrive shaft, and the signal was trans-mitted using all FM telemetry system.The torsional oscillation of the coolingfan were measured using an HBM tor-siograph.

The vibration data were gatheredover a two minute period as theengine speed was increased from 1200to 1800 rpm. During the tests thepump was operated with recirculatedwater. Speed rasters of the dynamictorque measured in the drive shaft forboth configurations are shown in Fig-ures 2a and 2b. The most significantengine harmonics from the speedrasters and the overall dynamic torque

(all harmonics combined) are plottedversus engine speed in Figures 3a and3b. For the original system, the overalldynamic torque in the drive shaft wasas much as 56% of the transmittedtorque. However, with the rubber cou-pling installed the overall dynamictorque was reduced to 16% of thetransmitted torque. The measuredtorque data compared favorably withthe predicted results from the comput-er analyses. For example, in Figure 2athe 4.5x engine harmonic passedthrough a resonance near 1400 rpm,and the amplitude of the 3.5x harmon-ic increased as the engine speedapproached 1800 rpm. The stress lev-els in the input gear shaft would bereduced by approximately the sameamount as the dynamic torque in the

drive shaft. The torsional oscillation atthe cooling fan was also reduced.Based on the test results, the rubbercoupling was recommended as a per-manent solution. No failures havebeen reported since the coupling wasinstalled.■

Troy Feese is a Project Engineer atEngineering Dynamics Incorporated inSan Antonio, Texas. He has six years ofexperience performing torsional and lat-eral critical speed analyses and rotor sta-bility analyses and in evaluatingstructural vibration problems using finiteelement methods. Mr. Feese is a memberof ASME, and the Vibration Institute,and he is a registered Professional Engi-neer in Texas.

Figure 2a. Original system – dynamictorque in drive shaft

Figure 2b. Modified system with rub-ber coupling – dynamic torque indrive shaft

Figure 3a. Original system – dynamictorque in drive shaft

Figure 3b. Modified system with rub-ber coupling – dynamic torque indrive shaft


A Noisy Problem

Two 60-inch condenser circu-lating water pumps, designedfor 126,000 gpm and 34 ft ofhead and driven by 1250 hps

195 rpm motors, were installed in anuclear generating station (Figure 1).Initially the performance of thesepumps was not acceptable eventhough rated flow was being deliv-ered. During normal operating condi-tions the pumps ran smooth, butthey occasionally produced rumblingnoises that were irregular and puz-zling. A typical 15-minute recordingrevealed 18 distinct rumbles fromone pump, 9 from the other. Thenoises were faint at first, grew to apeak, then gradually disappeared.These sounds occurred during singleand multiple pump operation.

It was initially believed thatthese occasional rumblings did notsignal a serious internal problem and

could remain an unsolvedcuriosity. One day, however, alake storm created large sur-face vortices in the forebay.Severe vibrations of thepumps and motors resulted.Operators would have shutdown the pumps if someonehad not started the travelingscreens, whereupon the rum-blings and vibrations quietedonly to resume after thescreen cleaning was stopped.High vibrations persisted dur-ing storm conditions and theproblem had to be resolved.

Type of Pumps UsedPumps for this service are usual-

ly vertical, wet-pit, mixed-flow dif-fuser types – a relatively inexpensivedesign that requires minimum space.However, such pumps are not veryaccessible for maintenance. Thepumps in this case were a vertical,

dry-pit design employing avolute casing rather than adiffuser and were moreaccessible for mainte-nance.

Both pump typesemploy a suction bell thatuniformly directs andaccelerates the water intothe impeller eye. For opti-mum performance it iscrucial that the suctionbells for these pumps beadequately submerged andcorrectly located in a prop-erly designed suction pit.

Initial DiagnosisThe pump manufac-

turer was confident that the noisewas not mechanical since it wasirregular and low in frequency andnot related to pump speed. The fore-bay water level was never belowdesign minimum of 12 ft, 6 in abovethe impeller eye, an elevation morethan adequate to prevent cavitation.If necessary, the pumps would havebeen able to pull a suction lift. Smallsurface vortices occasionallyappeared in the forebay around thedividing wall and stop log supports atthe entrance to the pump suctionchambers (Figure 2). However, noiseoccurred even when these vorticeswere not present. Surface vortices, ifstrong, can funnel air and cause thewater to prerotate as it enters thepump impeller. This produces noiseand less than optimum pump perfor-mance.

The only probable causesremaining to be considered by thepump manufacturer were air accu-mulating under the pump flooraround the outside of the suctionbell, air in the pump, or subsurface


HANDBOOK??


Modeling Pump Intake Noise

Huge circulating water pumps are “scaled-down” to solve a mystery.By J. P. Messina, Consultant

Motor

Pump

Screen

Maxiumlake level

Stop-logslot

Pump suction bell

Figure 1. Two 60 ″ vertical volute pumps and condenser circulating water intake

Pumpsuctionbell

Surfacevortex

Surfacevortex

Screen

Figure 2. Forebay surface vortices

vortices. A vent was installed torelease air by making a hole in thepump floor and adding a verticalpipe open at the top above highwater level in the forebay. However,no air was found. The pumps didemploy balancing rings on top of theimpeller – a low pressure area thatcould conceivably collect air. Thisarea was drilled into and vented, butthis did not take care of the problem.

Action TakenThe last possible cause of the

noise to be considered was underwa-ter vortices, but it could not be deter-mined if these indeed were occurringsince the area in question was hid-den under the pump floor. Becauseof the pumps’ size, large and costlybaffling would be required to correcta subsurface vortices problem, andat best it would probably be a hit andmiss procedure. Furthermore, therewas no proof that underwater vor-tices existed – or, if they did, thatthey were producing the noises.

Those involved decided to builda relatively inexpensive 1/10 scalemodel of one pump chamber (Figure3). Only the suction bell of the pumpwas modeled, and circulation ofwater was provided by an externalpump and stilling tank. The modelwas made of wood with Plexiglaswindows and measured 7 ft long, 4ft high and 2 ft wide. While the fore-bay, the suction bell and its chamber,and water depth dimensions wouldbe 1/10 actual size, what should themodel flow and velocity be? Thatwas a critical question.

The idea of modeling pumpintakes is very controversial.While it is important to main-tain geometric similaritybetween model and prototype,it is equally important to attaindynamic similarity. For a morein-depth explanation of model-ing laws, see the Pump Hand-book, section 10.2, 2nd edition,published by McGraw-Hill.

Since the pump intake is acombination of open and closedconduits, the model flow andvelocity should satisfy both theReynolds and Froude numberlaws. Unfortunately, at any setof test conditions both laws can-

not be satisfied at any one time. Ifthe model were tested to have thesame Reynolds number as the proto-type requiring 10 fps under thepump floor, very turbulent flowwould result in the forebay, prevent-ing the formation of surface vorticesthat solely or partly could contributeto a noise problem. If the intake weresolely an open pit, Froude similaritywould be the way to go. A more con-servative test favored by some inves-tigators for an open pit would be totest at the same velocity as the proto-type – or 1260 gpm model flow. Itwas therefore decided to test themodel for a flow range of 398 gpmFroude flow to 1260 gpm equal pro-totype velocity flow, not to a 12,600gpm Reynolds flow.

Model ObservationsIn the flow range tested no sur-

face vortices appeared. At a flow of

700 gpm, which is 1.7 times Froudevelocity and .55 times prototypevelocity, vortices appeared under thepump floor and around the suctionbell. They started horizontally fromeither side wall and were drawn upinto the pump suction bell. Theylooked like flashes of lightening,appearing one or two at a time (Fig-ure 4).

Figure 5 shows the test resultsand plots the number of vorticesappearing per minute vs. flow ingpm. It was interesting to note thatbetween Froude and prototypevelocities the vortex rate averagedabout 1 or 2 a minute, or 15 to 30 in15 minutes. The two prototypepumps were recorded producing 9and 18 rumbles in a similar 15minute period. Was this just a coinci-dence? The pump manufacturer didnot think so.


Figure 3. 1/10 scale model

Flow Meter

Stilling Tank

Suction BellWindow

Pump

ModelForebay

Axis of underwater vortex

Horizontalbaffle

Suction bellextension

Figure 4. Formation of underwatervortex and underwater baffling

10

9

8

7

6

5

4

3

2

1

0400 1200 2000 2800

Flow rate, gpm

Und

erw

ater

Vor

texi

ng r

ate,

no.

/min

. Froude velocity

Prototype velocity

Average oftest points

Figure 5. The number of vorticesforming per minute nearly equals the number of rumbles in model pump’stest velocity range.

Motor

Pump

Screen

Maxiumlake level

Stop-logslot

Pump suction bell

Figure 6. Formation of underwater vortex assuggested from model observations

Figures 4 and 6 show how theunderwater vortices formed and sug-gest where baffles might be addedand/or alterations made to the suc-tion pipe. Sixteen different modifica-tions were tried. Each was evaluatedbased on the number of underwatervortices formed per minute over anextended range of flows from 300 to2150 gpm.

The simplest and most effectivemodification was moving the backwall right up against the suction bell.This produced no underwater vor-tices. If this modification were to bemade in the field, however, it wasfeared that the suction bell would nolonger serve its original purpose,which was to direct flow to theimpeller eye evenly. The suction bellwould then act more like a shortradius elbow, hydraulically unbal-ancing the flow, and this could resultin rough running and less than opti-mum performance.

An equally effective test modifi-cation would have required extend-ing the suction bell to .4 belldiameters from the floor (presently 1diameter), moving the side and backwalls closer to the bell, i.e., 1.8 diam-eters width (presently 2.4 diameters)and 1 diameter from the bell center-line to the back wall (presently 1.5diameters). These modificationswould be more in keeping with therecommendations of the Hydraulic

Institute Standards (Fig-ure 7). Making thesechanges in the existingsuction chamber wallsin the field would nothave been practical.

RecommendedModifications

One modificationtested which showed nounderwater vortex for-mation and stillretained a good flowdistribution to thepump impeller waseventually recommend-ed to the customer. Thisrequired extending thesuction bell to half a diameter abovethe floor and placing a horizontalsteel plate at bell bottom elevation inthe rectangular area between the belland back wall (Figure 4). The steelbell extensions and plates were madein sections to facilitate installation inthe field. The horizontal plate wouldbe supported on columns.

A further recommendation wasthe addition of a steel, anti-surfacevortexing, horizontal surface plateand vertical skimming wall to beplaced between the stop log dividingwalls and opposite walls. This wouldbe attached to a gang of several stoplogs which, when lowered in place,would float on the surface as depict-ed in Figure 8. It was theorized thatduring a storm the three sets ofscreens could become overloadedand discharge unequally if left to col-lect screenings. This could create anunsymmetrical flow pattern into thepumps. The flow would thenapproach either of the stop log divid-ing walls at an angle and cause sur-face vortices. Also, these surfacevortices could feed into an underwa-ter vortex, increasing its strength andadding air to the pump. Because themodel only contained half of theforebay and screens, the unsymmet-rical flow due to an unevenly loadedscreen could not be observed causingsurface vortices.

ResultsThe recommendations were fol-

lowed, and the noises completely dis-

appeared. It can be concluded thatthe underwater vortices caused thenoises, and this was due to theimproper spacing of the suction bellrelative to the floor and walls. It wasalso concluded that noise can bemodeled in the range of prototypeand Froude velocities.■

J. P. Messina is a pump andhydraulics consultant in Springfield, NJand co-author of the Pump Handbookpublished by McGraw-Hill. He has pre-sented numerous papers on variouspump topics as well as taught courseson centrifugal pump theory, construc-tion and operation. He can be reachedat (973)379-5483.


1.5D

1.8D 2.4D

.4D

D

D

D

Figure 7. Installed vs. HydraulicInstitute suction pit dimensions

Pumpsuctionbell

Surfacevortex

Surfacevortex

Bafflingadded tostop surfacevortexing

Screen

Figure 8. Location of surface vortices and surfacebaffling

O n May 13, 1997 our compa-ny was called upon byColonial Pipeline to per-form a complete vibration

analysis on the circulating waterpump for its #2 cooling tower. Theunit had repeatedly experiencedmotor bearing failures caused byexcessive vibration. Detailed vibra-tion examination of the verticalpump, including spectral and phaseanalysis along with impact and coast-down testing, revealed that themotor vibration was caused by reso-nance. The motor speed of 720 rpmwas very close to a structural naturalfrequency, which was amplifying thevibration to very high levels. Thevibration was reduced from 1.4 inch-es/seconds (ips) to .06 ips by provid-ing additional stiffness to the top ofthe motor. This shifted the naturalfrequency to 827 CPM, which isabove the motor’s running speed,and thus solved the resonance prob-lem.

When we arrived at the site, weran the newly rebuilt motor uncou-pled to evaluate the severity of thevibration problem and to try todetermine the source of the vibra-tion. As shown in Table 1, the as-found vibration was indeedexcessive and highly directional. TheAlarm 1 level for a vertical pump ofthis size (12-20 ft) is .6 ips, and theAlarm 2 level is .9 ips. The level ofvibration at the top of the motor inthe horizontal direction was at 1.47ips, far exceeding the Alarm 2 level.For purpose of agreed-upon orienta-tion on a vertically mounted motor,the direction in line with the dis-charge piping is considered vertical,and the direction in line with themotor junction box is consideredhorizontal. The vertical vibrationreadings were much lower in ampli-tude than the horizontal, as shown inTable 1. If this was only an unbal-ance condition or some othermechanical defect, the vibration

would be similar in both directions.Having an amplitude ratio from onedirection to another of greater than3:1 is often an indication of a reso-nance problem. The phase data col-lected on the motor also indicated aresonance problem. Table 1 showsthat the motor was in phase from topto bottom in both directions, but thephase from vertical to horizontal wasout of phase by about 150 degrees.The phase data were taken from twovibration transducers mounted inperpendicular planes (90 degreesapart). If this were purely a mechan-ical defect, the most likely scenariowould have the phase from onedirection to another approaching 90degrees. A resonance problem allowsthe phase to shift.

After collecting the as-founddata, we shut the pump motor offand monitored the overall vibrationas the speed decreased. The ampli-tude versus speed plot is shown inFigure 1. Note that the vibration had


HANDBOOK??


Motor Bearing Failuresin Cooling Tower

Water PumpVibration analysis unlocks solution.

By Cliff Hammock, Technical Services Unlimited

Definitions

ips = inches per second, is a unitof measure of velocity. The veloci-ty is the rate of change of the dis-placement of a moving part.

CPM = cycles per minute, a mea-sure of frequency. Hertz (Hz)which is cycles per second (cps), isanother commonly used measureof frequency.

BEFORE Vertical Opp Horizontal Vertical HorizontalSTIFFNESS Drive End Opp Drive End Drive End Drive End

Amplitude (in/sec)

Phase (degrees)

1.47251

.42105

.62255

.23102

AFTER Vertical Opp Horizontal Vertical HorizontalSTIFFNESS Drive End Opp Drive End Drive End Drive End

Amplitude (in/sec)

Phase (degrees)

.0688

.1923

.0381

.0926

Table 1. Vibration amplitude and phase data – before and after modification

an absolute drop of over 1 ips over a23 rpm range of speed. This wasanother indication that resonancewas present. After the shutdown,impact testing was performed todetermine if a structural natural fre-quency was occurring at or near the

motor running speed.An impact test is per-formed – while theequipment is not run-ning – by striking themachine with a softobject such as a rub-ber hammer and mon-itoring the naturalvibration that occurs.The natural frequen-cies of vibration willshow up as peaks inthe spectrum. Thismay not be the mosttechnically correctmethod of determin-ing natural frequen-cies, but it is quickand has yielded verygood information onmany occasions. Wedid impact tests inboth vertical and hori-zontal directions, andthe resulting spectraare shown in Figures 2and 3. It is evidentthat the natural fre-quencies in the verti-cal directions areslightly higher thanthose in the horizontaldirection because ofadded stiffness in thedischarge piping. Alsonote that the couplingaccess area is in thehorizontal direction,which yields less stiff-ness. In both direc-tions the motor speedis very close to the nat-ural frequency, thusamplifying the vibra-tion.

The natural fre-quency of a piece ofequipment dependson the mass (weight)and the stiffness of thesystem. The amountby which resonanceamplifies vibration

depends on the damping of theequipment. The formula for the nat-ural frequency is:

Fn = √(k/m),

where k = stiffness and m = mass.

As you can see, the natural fre-quency can be increased by increas-ing the stiffness or decreasing themass, or reduced by increasing themass (weight) or decreasing the stiff-ness. In this case it would be difficultto add mass to the structure ordecrease the stiffness, but additionalstiffness could be added to the top ofthe motor.

The motor will always havesome inherent vibration due to anyunbalance associated with the rotor.In the situation with the cooling tow-er water pump, that unbalance, nomatter how small, was being ampli-fied by the fact that the motor was ina resonant condition. If the motorspeed could have changed, therewould no longer have been a forcingfrequency to amplify the natural fre-quency. However, for this applica-tion changing the speed was highlyimpractical.

In an attempt to move the natur-al frequency above the motor speed,we decided to add stiffness to the topof the motor. This is preferable tooperating machinery below a struc-tural natural frequency, so that each


Overall Vibration vs SpeedRepresentation of Actual Data

Overall Vibration

15

1

0.5

0300 360 420 480 540 600 660 690 705 720

Speed (RPM)

Figure 1. Amplitude vs. speed taken during coastdown

MED - Colonial Pipeline-Clg Twr #2C11-P002 -V01 Motor Outboard Bearing

Frequency in CPM

PK

Vel

oci

ty in

In/S

ec.

Freq: 690.00Ordr: .958Spec: .04160

PK = .0974LOAD = 100.0RPM= 720.RPS= 12.00

Analyze Spectrum13-MAY-97 09:30

0.07

0.06

0.05

0.04

702.

67

0.03

0.02

0.01

0

0 1000 2000 3000 4000 5000 6000

Figure 2. Impact test on motor outboard bearing inthe vertical direction

MED - Colonial Pipeline-Clg Twr #2C11-P002 -H01 Motor Outboard Bearing

Frequency in CPM

Freq: 677.85Ordr: .941Spec: .04115

PK = .0745LOAD = 100.0RPM= 720.RPS= 12.00

Analyze Spectrum13-MAY-97 09:32

677.

85

0 1000 2000 3000 4000 5000 6000

PK

Vel

oci

ty in

In/S

ec.

0.05

0.04

0.03

0.02

0.01

0

Figure 3. Impact test on motor outboard bearing inthe horizontal direction

Photo 1. The jacking screw arrange-ment was designed so that the stiff-ness could easily be controlled andoptimized in both vertical and hori-zontal directions.

Photo 2. Close-up of jacking screw– a relatively simple and low costsolution.

time the motor is started, the motorspeed will not pass the natural fre-quency range, thus amplifying vibra-tion during each startup andshutdown.

The solution was a jackingscrew arrangement designed so thatthe stiffness could easily be con-trolled and optimized in both verticaland horizontal directions (Photos 1and 2). The results of impact testingfollowing the stiffening is shown inFigure 4. The natural frequency wasshifted in both the vertical and hori-zontal directions. The horizontaldirection natural frequency wasshifted from 677 CPM to 827 CPM,which is greater than 10% above themotor speed. The vertical directionwas shifted from 702 CPM to 765CPM.

The before and after stiffeningvibration levels are shown in Figure5. There was a dramatic decrease invibration in the horizontal direction,from 1.4 ips to .06 ips. The vibrationphase data also significantlychanged. The phase from top to bot-tom in still in phase, but the phasedifference from vertical to horizontalis now 65 degrees. This is much clos-er to the expected 90 degree phaseshift than the as-found 150 degreephase shift.

Because resonance is commonon vertically mounted machinery, Ihave seen similar situations withother vertical pumps. Although thiswas a significant problem that result-ed in numerous costly failed motorbearings in the past few years, the fixwas relatively simple and inexpen-

sive. While this particular solutionmay not be the answer for all verti-cal pump problems, it does showthat, with effective use of technolo-gy, answers can be had even for longterm problems.■

Cliff Hammock is President ofTechnical Services Unlimited, Inc., inVidalia, Georgia, a company specializ-ing in maintenance engineering consult-ing services, including vibrationanalysis, laser shaft alignment and bal-ancing. He is certified by the VibrationInstitute as a Vibration Specialist IIand is Chairman of the Georgia Chap-ter of the Vibration Institute.



Frequency in CPM

PK

Vel

oci

ty in

In/S

ec.

Freq: 840.00Ordr: 1.167Spec: .04350

PK = .0795LOAD = 100.0RPM= 720.RPS= 12.00

Analyze Spectrum14-MAY-97 15.19

0.06

0.05

0.04

827.

14

0.03

0.02

0.01

0

0 1000 2000 3000 4000 5000 6000

Figure 4. Impact test results after adding stiffness in thehorizontal direction


Frequency in CPM

PK

Vel

oci

ty in

In/S

ec.

13-MAY-97 08:48Freq: 713.08Ordr: 1.000Spec: .05516

713.

08

0 600 1200 1800 2400 3000 3600

Max Amp1.22

PlotScale

1.4

0

11-JUL-97 09:25

Figure 5. Vibration before and after stiffening

T his article takes a differentapproach to the subject ofpump cavitation. Ratherthan focusing on the techni-

cal theory of fluid in motion, it offersa hands-on explanation of ways forusers to determine whether or notthey have a cavitation problem andhow to find a practical solution. Sim-ple formulas and definitions arefound at the end of the article to aidfurther system testing.

What Is Cavitation?Cavitation is the formation of

partial vacuums in a flowing liquidas a result of the separation of itsparts. There are two types. Suctionside cavitation is by far the mostcommon form (probably 90% of allknown events). Discharge cavitationis significantly less common.

Suction side cavitation is arestriction on the suction side of thepump system which does not allowsufficient fluid to enter the pumpand be discharged. The pump reactsto pressure on the discharge side andproduces a higher flow of liquid thancan be drawn in on the suction side.Suction side restrictions or atmos-pheric pressure decreases the flow tothe pump, particularly in suction liftapplications. The pump produces ahigher flow of liquid than can be sup-plied to it, due to suction side restric-tions.

Discharge side cavitation is arestriction on the discharge side ofthe pump system which constrictsthe fluid flow out of the pump. Sinceliquid can’t escape, it is recirculatedin the pump casing, damaging theouter edge of the impeller and thecasing, or the casing ring if it is pre-sent.

A common cause of suction side

cavitation relates to the vapor pres-sure of the liquid. Liquids boil at spe-cific temperature and pressurepoints. For instance, we know thatwater will boil at 212ºF at sea level.Carbon tetrachloride boils at 170ºF.Benzene boils at 176º F. Dowthermwill not boil or vaporize until itreaches 494.3ºF. Ethylbromide boilsat 101ºF, etc. When liquids turn to agas (boil), they will cause cavitationin a pump. Solvent transfers fromoutside tanks often can become aproblem in the summer when ambi-ent temperatures heat the liquids totheir critical vapor pressures. The liq-uid doesn’t have to boil to be a prob-lem. If it gets close to boiling and thesupply to the pump is restricted, cav-itation can be, and often is, theresult.

How to RecognizeCavitation

Cavitation is relatively easy torecognize. In its mildest form, it pro-duces a sharp pinging noise that hasoften been likened to the sound ofcorn kernels or gravel going throughthe pump. If you suspect cavitationin your pump system but are notsure because you don’t hear thatnoise, put the blade end of a screw-driver on the pump casing and thehandle end up to your ear. This willenhance your ability to hear anysuch noise inside the pump.

Another sign of cavitation is thatthe discharge pressure gauge on thepump system will fluctuate wildlyover a 5-to-10 psi range at a high rateof speed indicating uneven dischargeflow. One must be careful to put anew gauge on the system and checkthe gauge tap opening to ensure thegauge is operating correctly. A prop-erly operating system will give a

steady pressure gauge reading withlittle or no variation during pumpoperation.

Cavitation has numerous unde-sirable side effects. Because thepump is not operating in its properhydraulic balance, it is subject tointernal stresses that cause shaftdeflection and premature bearingand seal wear. These are two othersymptoms of cavitation. If bearingsand seals constantly are beingreplaced in a particular pumping sys-tem, severe misalignment or cavita-tion is the probable cause.

What Causes Cavitation?The five most common reasons

for cavitation are:

1. The pump was oversized by thespecifying engineer or pump sales-person. Oversizing the pump occursbecause the specifying person doesnot conduct a detailed system analy-sis to determine the proper headpressure and flows required for theapplication. Even when calculationsare performed, there may be a ten-dency to “fudge” the numbers to be“safe.” In actuality, when the pumpis first started up, the discharge pipesare new; therefore, the losses in thesystem are less than originally calcu-lated. The resultant pump oversizingis the most common cause of cavita-tion.

2. The second most common reasonfor cavitation is a change in the sys-tem demands. This can be illustratedby a spray system where a givennumber of nozzles is used and theback pressure against the pump toforce water through the nozzles atthe desired flow rate would be 100lbs. At 100 lbs., this theoretical pumpmay discharge 100 gallons per


HANDBOOK??


Cavitation in a NutshellHere’s a simplified approach to determining whether you have

a cavitation problem and what you can do about it.By Jeff Hawks, Buckeye Pumps, Inc.

minute on the performance curve.As the nozzles wear out, the open-ings through which the water passesare eroded. More water is allowed toflow through the nozzles, thus lower-ing the head pressure against thepump. The pump attempts to pumpmore and more liquid, but supplycan’t keep up with demand. Now thepumping system produces only 50lbs. of pressure at the discharge ofthe pump, and the flow through thepump, depending on the characteris-tic shape of the centrifugal perfor-mance curve, may be 300 to 500gallons per minute. The pump is nolonger operating in its best efficiencyrange due to a change in the systemperformance requirements, whichmay very well appear one day seem-ingly “out of the blue.” A commoncomplaint is the pump was workingfine yesterday, running well foryears, and suddenly it begins to cavi-tate!

3. The third most common cause ofsuction side cavitation is on a suctionlift or a pump whose suction sidesupply comes from a pit below thecenterline of the pump. In this situa-tion, debris within the pump canblock the suction and restrict theappropriate amount of fluid it needsto operate at peak efficiency. Also,leaks can develop in the suction line,and air is thereby introduced into thepump.

4. As stated earlier, temperaturecombined with marginal suction sup-ply can cause cavitation. Changes inthe process or unusual swings inatmospheric conditions are the mostcommonly observed reasons.

5. Lastly, as discharge lines in thesystem corrode or plug, pump dis-charge output is restricted, and dis-charge cavitation can occur. Checkvalves not operating properly oneither the pump discharge or suctionside also contribute to a state of cavi-tation.

How to Verify CavitationBeyond the obvious and charac-

teristic noise described earlier andthe erratic discharge pressure gauge,an inspection of the impeller in acentrifugal pump also will reveal the

effects of cavitation. It is importantto note that under proper operatingcircumstances impellers simply donot wear out. If it appears as if “ironworms” have eaten through the cen-ter of an impeller, there is suctionside cavitation. If you notice damagearound the outer diameter of thepump impeller, and in the casing,this is probably evidence of dis-charge cavitation.

Cavitation is not unique to cen-trifugal pumps. Cavitation is the for-mation of partial vacuums – orbubbles – in a flowing liquid as aresult of the separation of its parts.When these partial vacuums col-lapse, they pit or damage parts ofwhatever they contact, particularlythe metal surfaces or the elastomericsurfaces of a pump. In other words,cavitation affects every type of pump— centrifugal, progressing cavity,gear pumps, sliding vane pumps, air-operated diaphragm pumps — or anyother device that applies energy tofluid. The laws of physics apply to allpumps and to all systems.

To prove cavitation, install acombination gauge (one that reads invacuum and psi) on the suction sideof the pump and a discharge gaugeon the discharge side of the pumpand take the readings. The dischargepressure, plus suction pressure orvacuum, will be the operating pres-sure at which the pump is perform-ing. To avoid doing extensivecalculations, assume that 1 inch ofmercury on the vacuum gaugeequals 1.33″ of head, and rememberthat 1 psi equals 2.31 feet of head.(Centrifugal pump curves measuredischarge in feet of head, NOT inpsi.)

To illustrate, if there is a suctionlift condition and a vacuum pressurereading of 5 inches of mercury onthe suction combination gauge, con-vert that to 5 feet of head. If the dis-charge pressure gauge reads 100 lbs,multiply that by 2.31 and see that thedischarge pressure in feet of head is231 feet. Add the 5 feet of suctionhead to 231 feet, and you will findthat the pump is operating at 236 feetof head. If there is a positive headcondition on the suction gauge and itreads plus 10 psi, multiply 10 x 2.31.The result is 23.10 feet. Deduct that

23.10 feet of head from the dischargepressure of 231 feet of head, and youwill determine that the net pumpoperating point is 207.90 feet ofhead. Refer to the rotating shaftspeed of the pump to find thepump’s operating performancecurve. Then determine from thesereadings where the pump is operat-ing on its performance curve. Makesure that the pump performancecurve matches the motor speed.Motors can be switched from onerpm to another.

To refine these measurementsfurther, take amp readings on themotor inlet leads and convert themto brake horsepower. This willenable you to pinpoint the horsepow-er performance on the pump curve,which will be a double-check on thepressure readings taken earlier.

How to Temporarily Correcta Cavitation Problem

This is fairly easy for suctionside cavitation. If there is a valve onthe pump’s discharge side (and thereshould be), close the valve slowlyuntil the cavitation noise disappears.Conversely, if opening the valve tofull open makes the noise disappear,this probably is discharge cavitation.Other restrictions downstream maycause the problem to continue evenwith the valve open.

One may think closing the valvewill restrict the flow of liquid to thesystem. In reality, the system isfilled with fluid separated into gasbubbles, which may restrict the fullflow of liquid which you think youare getting. By returning the pump toits correct operating condition, youproduce a steady stream of gas-freefluid and this will render the mostefficient flow of material from thepump that can be expected under itscurrent operating conditions.

In the case of discharge cavita-tion, it may be necessary to recircu-late some of the liquid from thedischarge side of the pump back tothe supply of liquid. Do not recircu-late liquid directly to the pump’s suc-tion side as this will not alleviate theproblem. With discharge cavitation,it is necessary to bypass some of thefluid out of the discharge line so that


the pump operates as if it is produc-ing more flow than it really is. Fluidflow will continue, but now thepump will stop self-destructing.

While the above temporary fixeswill work for centrifugal pumps,they will not work as well for positive displacement pumps. Any temporary corrective action for posi-tive displacement pump cavitationshould only be done by an experi-enced pump technician or field engi-neer.

A Word of Warning: Neverrestrict the flow on the dischargeside of a positive displacementpump because it can cause per-sonal injury or damage thepumping system.

How to Correct CavitationSo It Does Not Reoccur

The only way cavitation can beeliminated is to analyze your systemprecisely and determine head pres-sure and flow requirements. Thisevaluation produces a system headcurve that can be used to determinethe correct size and type of pump to

do the job. Many times merely trim-ming the impeller or changing thespeed of the pump corrects the prob-lem. Perhaps 50% of the time it isnecessary to replace the pump withone more properly suited for theexisting system. The remaining cavi-tation problems can be corrected byaltering piping and/or supply eleva-tion, or by temperature regulation.System changes can be made, includ-ing cleaning out the pipes, removingobstructions, or replacing worn com-ponents. These measures will solvethe cavitation problem with little orno expense. One should also consid-er overall system performance andtry to enhance the system when cor-recting the cause of the cavitation.The key is to reduce the overall costof operation by reducing mainte-nance costs and improving efficiencythrough proper equipment sizing.Identifying a cavitation problem,understanding the physical causesand knowing how to deal with theproblem can help minimize costlydowntime and optimize pump per-formance.

Helpful Formulas andDefinitions

Ambient Temperature The nor-mal temperature at any given loca-tion at any given time.

Atmospheric Pressure is 14.7psi or 33.9 feet of water under stan-dard conditions at sea level.

Implosion The collapse orinward bursting of a bubble.

The Net Positive Suction Head(NPSH) The total suction head infeet of liquid (absolute at the pumpcenterline or impeller eye) less theabsolute vapor pressure (in feet) ofthe liquid being pumped.

Vapor Pressure The pressure atwhich liquid will begin to vaporize.This pressure is relative to the tem-perature of the liquid.■

Acknowledgments for contribu-tions to this article: Steve Cooper, DickBonesteel and Ed Plummer, BuckeyePumps Inc.; Andy Fraher, Flygt Pumps;David Doty, Moyno Industrial Prod-ucts; and Dana Maselli, Ingersoll-Dresser Pumps.



HANDBOOK?


Care must be taken in thedesign of a typical municipalpumping station or similarapplications to ensure that

the piping system does not imposeexcessive loads on the pumps. Exces-sive loads create the potential forincreased maintenance expense dur-ing the life of a pump. They also canincrease downtime and reduce pumplife expectancy. Since the purpose ofa pump station is to house the pumpand associated equipment, which, alltotaled, may represent a very signifi-cant investment, conventional wis-dom dictates providing anenvironment as close to ideal as pos-sible for the pump. This will helpachieve the best possible return onthe investment in the total structure.

A station designer following gen-erally accepted good practice wouldnot plan to load a pump with forcesapproaching those associated with,and best restrained, by pipe anchors.Yet improper attention to the pipingdesign in the vicinity of the pumpcan cause pump loads approachingthis magnitude and in effect use thepump as a pipe anchor – a situationthat is far from ideal.

This article is intended to alertreaders to an often overlooked aspectof piping design that can create loadsof pipe-anchor magnitude on pumps.That consideration is the hydraulicpressure reaction force of a pipingsystem.

Stretching and PullingSimply stated, all piping will

stretch some due to pressure in thepipe. The amount of stretch becomessignificant in a long or axially flexiblepipe. When such a pipe is connected

to a pump, and the other end of thepipe is restrained or resists this elon-gation in some way, the hydraulicpressure reaction force on the pumpmay be excessive. The forces pro-duced approach those intended forpipe anchors and can be strongenough to move the pump out ofalignment, overcome internal run-ning clearances, and excessively loadthe pump casing, base and anchorbolts. A shear load and moment willalso be transmitted to the pumpfoundation and the station floor.

As can be seen, failure to consid-er this effect can create havoc andresult in an unsatisfactory installa-tion – one that is both detrimental tothe life of the pump and unsatisfacto-ry to the owner. This is one reasonwhy the recommendation is provid-ed in the current Hydraulic InstituteStandards ANSI/HI 1.1-1.5-1994 forcentrifugal pumps, page 119, as fol-lows: “Suction and discharge pipingmust be anchored, supported, andrestrained near the pump to avoidapplication of forces and moments tothe pump.” A similar statement isfound in ANSI/Hydraulic InstituteStandards ANSI/HI 2.1-2.5-1994 forvertical pumps, page 65.

Complications fromExpansion Joints andFlexible Pipe Couplings

An extreme case of pump load-ing due to a hydraulic pressure reac-tion occurs when an expansion jointor flexible pipe coupling is usedbetween a pump and an anchor orsome other restraint with no tie rodsto restrain the resulting force. In thiscase, the force on the pump will be

equal to the pressure times the pro-jected area of the maximum insidediameter. As the standards state,“This force may be larger than can besafely absorbed by the pump or itssupport system.”

In most cases the force will belarger than can be safely absorbed bythe pump because the pump typical-ly lacks the mass and strength of apipe anchor that is designed torestrain such forces. Consider thefact that a pump must to a certaindegree be “open” or “hollow” withindue to the hydraulic passagewaysthrough which water must pass inorder for the pump to pump!

Another factor that can make theproblem even worse is when the dis-charge pipe has been made largerthan the pump discharge by using anincreaser (not a bad practice initself), so that the pipe coupling orexpansion joint is on a larger pipe.This creates a larger force on thepump. By not keeping such loads offthe pump, not only is a relativelyvaluable machine being heavilyloaded unnecessarily, natural reac-tions to those loads are produced inthe pump foundation and pump sta-tion floor. Again, reference is madeto ANSI/Hydraulic Institute Stan-dards, which recommend that a pipeanchor be installed between an axial-ly flexible pipe coupling or expansionjoint and the pump to absorb the axi-al force. In cases where properanchoring cannot be provided, theuse of adequate tie rods to protectthe pump and expansion joint or pipecoupling is acceptable (as is also stat-ed by the ANSI/Hydraulic InstituteStandards).

Hydraulic Pressure Reactionsin Pump Piping Systems

By Jack Claxton, Patterson Pump Company


Tie Rod ProblemsThis brings us to a second case of

excessive loading due to thehydraulic pressure reaction – loadscaused by poorly designed tie rods.In determining the adequacy of tierods, axial deflection as well as stressmust be evaluated. A common mis-take is to size the tie rods based onallowable stress without consideringdeflection. High strength steel isoften used for the tie rods, and a cor-respondingly high value of allowablestress is used to determine the sizeand number of rods. If this is thecase, a relatively high value of deflec-tion will occur because the modulusof elasticity that affects deflectionremains essentially the same as thatof carbon steel. Rods so designed willbe relatively flexible axially com-pared to the forces imposed uponthem. While such designs may beacceptable in many instances, whenthey are used near a pump that mustremain aligned with its driver withinthousandths of an inch, the result isan unacceptably high reaction forceon the pump.

As a short cut to analyzing theforces and deflections involved insuch an application, a specifyingengineer might decide to use tie rodsbased on a standard that does notconsider axial deflection and itseffect on pumps. Many design firmsuse internal company standards thatmay be inadequate in this way. Oftentie rod designs encountered inmunicipal or similar pumping appli-cations comply with or are compara-ble to American Water WorksAssociation (AWWA) standards.These standards (AWWA M11, ThirdEdition, for example) are suitable formany applications in whichhydraulic thrust forces must berestrained. But in applications nearpumps, the 40,000 psi allowablestress design criteria upon whichthey are based allows unsuitabledeflection values near a pump with-out a pipe anchor or other means torestrain the pump side of the piping.In addition to the high allowablestresses, the length of the tie rods

that affects axial deflection is notspecified. Tie rods used near pumpsare best kept as short as possible.

Furthermore, axial deflection is afunction of the cross sectional area ofthe tie rod or pipe. The cross section-al area of the recommended tie roddesign for a given pressure and pipesize is found (in AWWA MII) to beapproximately one-fourth to one-third of the corresponding recom-mended pipe. This inherentlyintroduces an axial flexibility at thetie rod location, as compared to solidpipe. The point is not whether or notto use such standards in the design ofpiping systems, but to verify that thepiping system will not place exces-sive loads on the pump in any givendesign.

A Case in PointTo illustrate the problem, one

pump installation involved a 24”sleeve type pipe coupling with four1-1/8” diameter tie rods 46-1/2” long.The pipe was connected to a mani-fold that restrained the axial deflec-tion of the 24” pipe at that end. Thepipe pressure of 246 psi produced athrust force of approximately 55tons. The calculated axial deflectionacross the rods was only 0.043” butactually measured 0.72” using dialindicators. This was most likely dueto the bending of the lugs that heldthe rods. The adjacent double suctionpump, which fortunately had not yetbeen doweled to its base, moved outof alignment with its motor, resultingin overheated inboard pump andmotor bearings.

I say “fortunately” here becausehad this load been restrained by thedoweled pump and the pump hadsomehow got through the start-upperiod without the problem beingmanifested, the tremendous forcesimposed upon it most likely wouldhave had an adverse effect at sometime.

Alternative SolutionsOne approximation to avoid

excessive tie rod flexibility is todesign rods to have the same axial

rigidity of that of the piping. Thetotal tie rod area that provides equiv-alent rigidity is AR=Ap (Ep/ER),where AR = total tie rod cross sec-tional area, in2, Ap = cross sectionalarea of the pipe metal, in2, Ep = themodulus of elasticity of the pipematerial, and ER = the modulus ofelasticity of the tie rod material. If tierod material and pipe rigidity are thesame, it is a simple matter to deter-mine the approximate length of pipeto the nearest pipe anchor for a givenvalue of allowable deflection. In thecase of relatively rigid pipe, it maynot be practical to obtain equiva-lence. In this instance, using as manyrods as practical is advised, keepingthem as short as possible to limittheir effect.

Another approach is to limit theallowable value of the tie rod deflec-tion to a very small value – say0.005”. This may be difficult, but0.005” is comparable to allowablehorizontal misalignment values of apump with its driver.

Once the tie rod design has beenestablished, the entire run of pipe tothe nearest anchor can then be ana-lyzed to verify that the deflection isnot excessive. (Short runs of pipingare best.) Using pump and baseplaterigidity information provided by thepump manufacturer, calculations canthen be made to ensure that the man-ufacturer’s maximum recommendednozzle loads are not exceeded.

Other aspects of piping designthat can adversely affect a pump andthat are beyond the scope of this arti-cle will need to be considered also.These include piping misalignment,thermal expansion and contraction,and the weight of the piping and itscontents.■

Jack Claxton is Vice President ofEngineering for Patterson Pump Com-pany in Toccoa, Georgia. He has 23years of experience in hydraulic andmechanical design, as well as applica-tion and field troubleshooting for verti-cal and centrifugal pumps. He is agraduate of Georgia Tech and is activein the Hydraulic Institute.


Pipe Anchor Expansion Joint orFlexible PipeCoupling(no tie rods)

FLOW

ARRANGEMENT 1: NOT RECOMMENDEDAn expansion joint or flexible pipe coupling between the pump and the nearest anchorallows a force equal to the pressure in the pipe times the area corresponding to the max-imum inside diameter to be put on the pump. This force is transmitted from the nozzle tothe casing, pump tie-down bolts, base, anchor bolts, pup foundation, and pump stationfloor. This arrangement is not recommended by Hydraulic Institute and pump authori-ties, because it is impractical to design pumps to withstand this force and the pump willessentially be used as a pipe anchor. This arrangement can produce reaction forces ofsuch magnitude to cause catastrophic failure or reduced pump life, and is therefore notrecommended.

ARRANGEMENT 2:Better than Arrangement 1 but probably not as good as Arrangement 3 because of thelikely use of undersized tie rods. Tie rods are frequently designed for an adequate safetyfactor considering stress and using alloy steel. The alloy steel allows higher designstresses, but gives no additional resistance to deflection compared to carbon steelbecause the modulus of elasticity of alloy steel is the same as that of carbon steel. Thissets up a potentially excessive axial deflection situation due to the axial flexibility intro-duced into the piping arrangement by the tie rods. For a tie rod arrangement to be equiv-alent to a rigid pipe arrangement (Arrangement 3) in terms of axial rigidity, the followingrelationship must be met:

AR = AP

AR = total cross sectional area of the tie rods (sq. in.)AP = total cross sectional area of the pipe material (sq. in.)ER = modulus of elasticity of the tie rod material (psi)EP = modulus of elasticity of the pipe material (psi)

ARRANGEMENT 3:Better than Arrangement 1 and probably better than Arrangement 2 depending on the tierod design but there still may be excessive axial flexibility in the pipe that will load thepump due to the long length of pipe between the pump and the nearest anchor. Whenthe pipe is pressurized, it will stretch and load the pump.

ARRANGEMENT 4: RECOMMENDEDRecommended as per ANSI/H.I. 1.1-1.5-1994 (centrifugal pumps) or ANSI/H.I. 2.1-2.5-1994 (vertical pumps). “Suction and discharge piping must be anchored, supported andrestrained near the pump to avoid application of forces and moments to the pump...”Also, “It is recommended that a pipe anchor be installed between an expansion joint andthe pump to absorb the axial force.” In addition, keep “L” as small as possible.

ARRANGEMENT 5: ALTERNATIVE DESIGNPer ANSI/H.I. 1.1-1.5-1994 and ANSI/H.I. 2.1-2.5-1994 “When proper anchoring cannotbe provided, adequate tie rods must be provided and properly adjusted to protect thepump and the expansion joint. Limit tie rod axial deflection to 0.005” for best results,and verify that the total axial deflection is not excessive.

Pipe Anchor

Expansion Joint orFlexible Pipe Coupling

with inadequate tie rods

FLOW

Pipe Anchor

No expansion joint orflexible pipe coupling

between pump & pipe anchor.

FLOW

FLOW

Pipe Anchor

L

If an expansion jointor a flexible pipe coupling is desired, place it on this sideof the pipe anchor

Pipe Anchor Expansion Joint or

Flexible Pipe Couplingwith adequate tie rods

FLOW

CLARIFICATION - PIPE SUPPORTShown above are typical pipe supports which support the weight of thepipe and its contents but do not restrain the pipe along its axis. Supportsof this type therefore are not to be considered as anchors as recommend-ed by the Hydraulic Institute Standards.

NOTES:

1. For the sake of illustration, a horizontal double suction pump isdepicted. The same principles will apply to other pump types.2. For the sake of simplicity, only the discharge side is shown,although the same principles apply to the suction side. There willtend to be more problems on the discharge side in axially flexible pip-ing arrangements due to the higher pressures encountered.3. This drawing discusses piping arrangement recommendations toavoid problems caused by axial flexibility in the piping when subject-ed to pressure. Effects of misaligned or offset piping and thermalexpansion are not discussed, which can also create pipe strain andthereby cause the pump to be subjected to a reaction force.

EPER

( (

Figure 1. Piping arrangement recommendations.


The purposes of collecting fail-ure data are to find problempumps that need to beaddressed, to trend improve-

ments and to compare results againstthose in other facilities. The body ofwork order data collected over sever-al years, even in a small facility, willbe large and varied. At the sametime, comparison with other facili-ties is difficult because there are somany different ways of reportingfailure information. The approachwe tried was to combine similar fail-ures together in categories. Thisenabled us to analyze a large numberof failures at one time, it simplifiedsome of the calculations required todetermine Mean Time Between Fail-ures, and it forced us to define fail-ures in detail, which should facilitatebetter comparisons to other facilities.The information collected is summa-rized in Tables 1 and 2.

Our facility is a small refinerywith a capacity of about 57,000 bpd.We have the standard process units:crude/vacuum, FCC, alkylation,reformer, hydrotreater, coker and iso-merization. And the refinery has theusual supporting units: boilerhouse,waste water treatment and tankfarm.

The most common type of pumpin our facility is the horizontal singlestage overhung impeller pump. Itaccounts for approximately 60% ofthe pump population. We do nothave an accurate count of meteringpumps. Therefore, they are notincluded in the pump count. Theaverage driver-rated horsepower is72.

Data CollectionThe failure data presented here

were gathered from the Work OrderModule in our Computer Mainte-nance Management System (CMMS).Since our CMMS was installed in

October of 1993, the informationcovers the years 1994 through 1997.All pump repairs are handled by per-sons in the machinist craft. Opera-tors generate the work orders for thepumps in their unit and tag the work


HANDBOOK?

Categories

Seal FailuresBall/Roller Bearing FailuresOverhaul of Packed PumpVertical Pump FailuresGeared High Speed PumpsPositive Displacement PumpsAll FailuresCoupling FailuresPacking Adjustment or Re-Packing

1994

2.59.75.22.62.22.31.7

12.71.0

1995

2.38.8

16.82.15.25.71.6

14.31.2

No. of pumps

346290(1)67462634449(2)34267

1996

3.47.8

13.41.84.3

11.31.9

14.31.6

Notes:(1) Vertical In-line pumps, geared pumps were not included.(2) Metering pumps, submersible pumps were not included.

Developing MeaningfulPump Failure Data

Improving performance by categorizing failure data: one refinery’s success story.By Oleh Berezowskyj, Clark Refining and Marketing

Table 2. Mean time between failures

1997

3.013.233.51.93.76.82.19.51.7

Pump Failure Categories

Seal FailuresBall/Roller Bearing FailuresCase Gasket LeakOverhaul of Packed PumpMaterial in PumpCorrosion/ErosionInternal RubbingSleeve Bearing FailureInfrequentVertical Pump FailuresGeared High Speed PumpsPositive Displacement PumpsHigh VibrationTotal

Non-Pump & Non-Failure CategoriesCoupling FailuresMinor RepairsPacking Adjustment or Re-PackingMetering (Controlled Volume) Pumps

1994

141308

13125600

1812154

264

199427196865

1995

151336464409

2256

26276

199524705477

1996

10237115

1510923

2563

13241

199624804259

1997

116226264143

2475

10210

199736303968

Table 1. Pump failure categories


orders for the machinists. Eachpump work order was reviewed byanalyzing the repair or discussing thework done with the machinists.

A work order was not counted asa failure until it was listed as com-pleted in the Work Order Module.The completed date on the workorder was then used as the failuredate. This way the failure was ana-lyzed and the corresponding man-hours and material costs werereviewed before the work order wascategorized. Furthermore, the previ-ous month’s data did not have to beupdated if the repair extended intothe next month.

We do not have standing workorders for minor repairs. A separatework order is written for most jobs ina unit, including adjusting packingand setting bearing housing oilers.Also, operators will generate workorders to investigate problems. Bothof these factors substantially increasethe number of work orders in ourCMMS.

The information in Table 1 waslimited to items under the control ofthe machinists craft. This was donefor two reasons. First, trackingrepairs by another craft adds sub-stantially more time and effort to fail-ure analysis. Second, some limitshould be set for the types of repairsthat are counted as part of a pump. Ifone includes motors, should the databe limited to bearing failures? Orshould problems with windings,start/stop switches, switch gear andwiring be included? Should piping,control valves and suction vesselrepairs also be counted as pump fail-ures?

In our facility, electricians areresponsible for all the motor workincluding replacing bearings. Conse-quently, motor repairs and motorbearing failures (including motors forvertical in-line pumps) were notadded to Table 1. Submersiblepumps were not included in the fail-ure data either because the electricalcraft is responsible for repairingthese pumps.

Similarly, any work on the suc-tion and discharge piping connectedto a pump, including cleaning suc-tion screens and repairing checkvalves, is handled by another craft

and was not included in this data. About 20% of the work orders

were not considered repairs or fail-ures and were not included in Table1. They were:

• Situations that were investi-gated and no problems were found.

• Work orders involving a prob-lem unrelated to the pump. An example of this type of work order is a performance prob-lem caused by operational dif-ficulties or a bad check valve. Even if the pump was disas-sembled, the repair was not counted if there was no prob-lem with the pump.

• Duplicate work orders.• Work orders generated to buy

parts. • Work orders generated for in-

house repairs of pumps or parts carried in warehouse stock. The original work order to repair the pump was counted.

• Upgrades such as installation of larger diameter impellers and mechanical seal conver-sions.

• Preventative Maintenance(PM)work orders such as changing oil and greasing. PM and inspection work orders sched-uled during turnarounds were not included either.

Reworks were not counted asseparate failures. The pump had torun for at least a few days before wewould count the second repair as afailure. For example, if a seal leakedon start-up (the most common situa-tion), it was not counted as anotherseal failure but was considered partof the original work order.

CategoriesThe classification of failures in

Table 1 is by no means 100 percentaccurate. First of all, the evaluationof a failure in many cases is subjec-tive. For example a damaged bearingis found in a pump that is beingrepaired because of a leaking seal.Did the bad bearing cause the seal toleak? Or was the bearing damagecoincidental? If the damage wassevere enough, the repair was count-ed as a bearing failure.

Secondly, a failure can belong to

more than one category. A number ofour positive displacement pumpshave internal product lubricated anti-friction bearings. Should a bearingfailure in this type of pump be count-ed as a regular bearing failure or as afailure of the pump itself? Since thebearing is internal to the pump and isin a different environment than abearing in a centrifugal pump, wedecided to count it as a pump failure.

Finally, if the definition of a cate-gory is changed, then past datashould be reclassified. In our case theonly information we have regardinga repair is the work order descrip-tion, which is filled in by the opera-tor, and time and materials chargedto the work order. Since memoriesfade quickly, it becomes difficult toreclassify work orders once they aremore than a few weeks old. The datafor minor repairs shown in Table 1illustrate the problem. They show asignificant drop in work orders in1997. Initially any work order thatthe machinists investigated and didnot find a problem was counted as aminor repair. Later we stopped clas-sifying these work orders as repairsor failures and did not add them any-more to the minor repairs category.However, we did not go back andreclassify older work orders. Onmany of them it was impossible todetermine from the descriptionwhere the work order belonged.

Seal Failures: Repair done toreplace a leaking seal. If failed bear-ings or heavy rubbing was found,then the repair was not counted as aseal failure. This category has themost amount of causes. We havefound at least 10 separate problems.

Ball/Roller Bearing Failures: Re-pair because of vibration, noise, hotbearings or a locked rotor wheredamaged bearings were found. Abearing that was replaced becausevibration analysis indicated a prob-lem was counted as a failure even ifthe damage found on the bearing wasminor. It was assumed the damagewas the initial stage of a failure.

Most bearing failures could bedivided into two types: lubricationproblems and surface damage. Fail-ures from improper lubrication werethe most dramatic. The retainerswere cracked, smeared or missing


(Photo 1). Sometimes the rolling ele-ments and bearing races weredeformed and discolored from exces-sive heat (Photo 2). This modeaccounted for about three quarters ofall the failures. Most of the time wesuspected that low oil level was thecause, but this was confirmed in onlya few cases.

In the second type of failure thebearing had some surface damagewhich caused noise and vibration.The work orders for this failure weregenerated by the operators and byour vibration analysis program. The

damage included spalls and dents onraceways and marks on balls andretainers. The source of the damagewas not determined in many of thesefailures. The causes we were sure ofare dirt in the oil, corrosion of theraces and impacts from poor mount-ing practices.

One other type of failure wehave found is heavy spalling on bear-ings under excessive loads (Photo 3).This problem has occurred on onlyfour pumps, all manufactured in the1940s.

Case Gasket Leak: Repair thatrequired replacement of the case gas-ket or clean-up of the case gasket seal

area. About a third of all the failureshave occurred on pumps in hydroflu-oric acid service. The cause is possi-bly corrosion under the moneloverlay in the casing. Other possiblecauses are improper installation,damage to the gasket surface and cor-rosion of the gasket surface.

Overhaul of Packed Pump: Workorder for packing sleeve replacementor stuffing box repair to correctexcessive packing leakage. More than60% of the work orders were gener-ated for four pumps in boiler feedwater circulating service, which werecently converted to mechanicalseals.

Material in Pump: Work ordergenerated because of poor perfor-mance, locked rotor or high vibra-tions where material (metal parts,coke, etc.) was found in the impelleror suction screen or where theimpeller was plugged with depositsfrom the product (lime, salts, etc.)

Corrosion/Erosion: Work ordergenerated because of poor perfor-mance or leakage through the pumpcasing where heavy corrosion or ero-sion of the casing, impeller or wearrings was found (Photos 4, 5 and 6 ).Three occurrences of impeller dam-age due to cavitation were includedin this category.

Internal Rubbing: Work ordergenerated because of high vibrations,seal leakage, locked rotor or poorperformance where heavy rubbing ofwear rings and bushings was foundupon disassembly. Most of these fail-ures have occurred on impellerbetween bearing type pumps, singleand multistage. We suspect loss ofliquid in the pump is the most com-mon cause of these failures. A few ofthe failures include pumps in ouralkylation unit. These pumps havean unusual internal corrosion prob-lem that causes the case wear ring togrow into the impeller wear ring andrub.

Sleeve Bearing Failure: Repairwhere a sleeve bearing was replaceddue to babbitt wiping. We have onlyseven pumps with sleeve bearings.All but one of the failures haveoccurred on two pumps in the sameservice. We are still investigatingthese failures.

Infrequent: A category encom-

Photo 2. Extrusion of inner race on5307 thrust bearing caused by over-heating

Photo 3. Spalling of outer race on5309 bearing from thrust overload

Photos 4 and 5. Heavy erosion of caseand stuffing box from catalyst carry-over

Photo 6. Corrosion of impeller in sourwater service with heaviest attack atthe grain boundaries.

Photo 1. Smeared bronze retainerresulting from lack of lubrication


passing odd failures such as: • water jacket, water cooler or

pump casing fractured from freezing • cast iron pump casing frac-

tured from mishandling• failure of impeller bolting on a

overhung pump caused by chemicalattack

Vertical Pump Failures: Workorder generated because of vibrationor poor performance on a verticalpump. A vertical pump included inthis category is any with a long shaftand bushings. This includes sumppumps and vertical turbine pumps(both deep well and can type), butnot vertical in-line pumps. Seal fail-ures and corrosion/erosion of thepump were counted in their own cat-egories. However, failures caused byrestricted suction or material in theimpeller were still counted in thiscategory. These failures were catego-rized this way because damage to thebushings and shafts caused by inade-quate flow appeared to be commonin this type of pump.

Geared High Speed Pumps: Fail-ure of the gearbox bearings, gears oroil seal. Product seal leakage, corro-sion/erosion, material in the pumpand case gasket leaks were includedin their respective categories. If afailed gearbox was found on a workorder written to repair a leakingproduct seal, it was assumed the leakwas caused by the gearbox vibration.The repair was counted as a gearboxfailure.

Positive Displacement Pumps: Anyrepairs. This category combinedrotary and reciprocating pumpsbecause we have only four recipro-cating pumps. Repairs of leakingseals were counted in the seal fail-ures category, but failures of internalproduct lubricated roller bearingswere kept in this category. Most ofthe work orders for rotary pumpswere for performance problems. Allthe work orders for the reciprocatingpumps were generated for twopumps that are no longer in service.The problems listed on these workorders included performance, noise,packing leakage and oiler malfunc-tion.

High Vibration: Work order forhigh vibration generated by ourvibration analyst (and sometimes by

operations personnel). The purposeof this category was to count predic-tive maintenance type repairs inwhich work was done (alignment,balancing) before a failure occurred.Most of the repairs (80%) requiredcorrecting misalignment and loose-ness (usually loose bolting) and didnot require disassembly of the pump.Those that required disassembly —mostly unbalance and vane passvibration — were kept in this catego-ry if no obvious problems (internalrubbing, corrosion/erosion, materialin pump) were found. Work ordersgenerated due to vibration analysisindicating a bearing defect werealmost always counted in theball/roller bearing failure category.Whenever a bearing was cut apartand analyzed, some kind of imper-fection was found. The imperfectioncould be considered the initial stageof a failure. To be conservative, thesework orders were added to theball/roller bearing failure category.

Poor performance for centrifugalpumps is not listed as a categorybecause in all cases where a workorder was generated because of poorperformance one of the previous fail-ure modes listed in the above cate-gories was found to be causing theproblem. Replacement of wear ringsis not listed as a category eitherbecause all the replacements of wearrings to date were due to rubbing orcorrosion/erosion, and the failureswere counted in their respective cat-egories.

Non-Pump and Non-FailureCategories

Coupling Failures: Most of thecouplings in our facility are urethanedonut shaped elastomers. The mostcommon failure mode for these cou-plings is cracking because of materialdegradation (“aging”). The rest of thecouplings are gear, grid, metal diskand miscellaneous types of elas-tomers. The most common failuremode for gear and grid couplings islack of lubrication (Photo 5). A fewfailures have been due to misalign-ment. Since the predominant failuremodes are a function of the couplingcondition, we feel coupling failuresare independent of the driven anddriving equipment. Therefore, they

were not counted as pump failures.They can be added back to the fail-ure categories if deemed necessary.

Minor Repairs: Any work that didnot require substantial disassemblyof the pump and where the pumpwas operable was counted as a minorrepair. Minor repairs included adjust-ing or replacing oilers, tighteningbolting, checking oil rings, repairingflush or cooling water piping, chang-ing oil filters, replacing oil seals andadjusting or replacing belts.

Packing Adjustment or Re-Pack-ing: This type of repair, considerednormal wear, was not counted as afailure.

Metering (Controlled Volume)Pumps: Our metering pumps do nothave equipment numbers, and we donot maintain the same type of recordson them as we do on other pumps.Therefore, the work order informa-tion is not that accurate and was notincluded with the other pump failuredata. Table 1 lists the number of workorders generated to do some kind ofwork on metering pumps. Also, inour facility the pipe-fitters craft caninstall metering pumps. In caseswhere a work order for a meteringpump was assigned to the pipefitters,it would not be listed in the table.Portable air operated diaphragmpumps, which are handled the sameas metering pumps in our facility,were included in this category.

Mean Time Between Failure(MTBF)

Calculating traditional MTBF foran individual piece of equipment isrelatively easy. From equipment his-tory, take the first failure date andsubtract it from the last failure date.This is the total time span for all therecorded failures. Count the numberof failures and subtract 1. This is thenumber of operating periods of theequipment in the time span. Dividethe time span by the number operat-ing periods and you have MTBF.Note that MTBF cannot be calculatedfor equipment with zero or one fail-ure.

MTBF = Time Span (Last Failure Date - First Failure Date)

Number of Operating Periods (Number of Failures - 1)


A variation of the MTBF calcula-tion ignores the true time span.Instead, the time span of the equip-ment history database is used. Thedatabase time span is divided by thenumber of failures, not operatingperiods, to produce an approximateMTBF. This eliminates the need tofind the first and last failure dates.With this method MTBF cannot becalculated for equipment that hasnever failed.

MTBF = Equipment History Database Time SpanNumber of Failures

Number of FailuresMTBF calculation for a group of

pumps is somewhat more complicat-ed. A group can be all the pumps in aunit or facility, a particular type ofpump (centrifugal single stage over-hung) or a particular configuration ofa pump (pumps with seals). I haveused two methods of calculatingMTBF for a group.

The first method, which is theone presented in the tables, is not atrue MTBF. It is a number that repre-sents an average time between fail-ures of a component or pump typefor an average piece of equipment.This type of calculation is easy andallows evaluation of failures the firstyear information is collected.

Table 2 shows the MTBF for allthe pumps in our facility and theMTBF for the most significant failurecategories. MTBF was not calculatedfor categories with a small number offailures and large population such ascase gasket leaks and internal rub-bing. Using the seal failure categoryfor 1997 as an example:

MTBF = 346 pumps with seals/116seal failures per year = 3 years.

The second method calculatesthe MTBF for each pump and thendetermines the average of all theindividual MTBFs. Since pumps withzero failures do not have a MTBF, adummy failure is added to thosepumps to permit calculation of theaverage. For this type of calculationthe time span of the equipment histo-ry needs to be as great as possible inorder to reduce the number of addedfailures. Calculations for only one

year will not be accurate because alarge number of dummy failures willhave to be added to the database.This hinders calculating MTBF in thefirst year data are collected. More-over, trending MTBF requires usingfailures from the beginning of theequipment history. This muddies theeffect of recent improvements.

Also, the second method isinsensitive to problem pumps. As thenumber of failures for an individualpump increases, its MTBF decreasesapproaching zero. In a large popula-tion of pumps, a few pumps withvery low MTBF (high failure rates)will not affect the overall average.

Significantly different MTBF canbe produced depending upon whichmethod of calculation is chosen. Forour facility, using data from the lastfour years, the overall pump MTBFwas calculated individually and aver-aged. The result was 2.82 (15% addi-tional failures were added toeliminate zero failures).

On the other hand, the overallpump MTBF calculated using thetotal number of pumps divided bytotal number of failures divided bythe time span equaled 1.82.

An accurate count of pumps isnecessary to calculate MTBF correct-ly and to compare accurately againstother facilities. We inventoried thepumps in the whole refinery beforeevaluating the MTBF. In two otherfacilities we found that almost half ofthe pumps on the master pump listwere removed or abandoned inplace.

Total number of pumps in-stalled, not the number of pumps inoperation, was used to calculateMTBF. This way we did not have toaccount for spare pumps operating inparallel with main pumps and forintermittent services such as pumpsin the tank farm. Refineries and mostother facilities are structured thesame way — spare pumps for mostservices, single pumps for intermit-tent services. Thus, comparisonsbetween facilities should still beaccurate.

SummaryThe amount of repair data gener-

ated for pumps can be overwhelm-ing. Trying to allocate one’s time and

the company’s resources effectivelyrequires accurate information. Cate-gorizing failures can reveal enlight-ening patterns and highlight problemareas. From Tables 1 and 2 we canconclude that:

• Seal failures account for the ma-jority of the failures and deserve the attention that they receive.

• Improving lubrication will have a significant impact on pump failures.

• Vertical pumps are a problem area that needs to be addressed.Comparisons from year to year

in a particular category can readilyshow where improvement efforts,both equipment specific andplantwide, are making a difference.Table 1 shows a 30% reduction inseal failures from 1995 to 1996. Overthe past few years we have beentroubleshooting problem seals andimproving materials, designs andflushes for individual services. Also,we have involved our machinists incorrecting seal problems. Half of thereduction is due to seal modifica-tions, and we believe the other half isdue to higher quality repairs.

There are many theories aboutfailure reporting, and they all pro-vide useful information. However,standardization — or at least anunderstanding of how the failurereporting was done — is necessary sofacilities can compare against eachother. The use of agreed-upon cate-gories promotes development of thekind of detail needed to make mean-ingful comparisons. Moreover, withcategories one can account for differ-ences in failure data. For example,metering pumps are not included inthe MTBF for all failures because ourcount of these pumps is not accurate.With categories we can still compareourselves to other facilities becausewe can account for this difference inthe data.■

Oleh Berezowskyj has held variouspositions as Rotating Equipment Engi-neer and Supervisor in corporate engi-neering and in refinery maintenanceduring the past 23 years. He is current-ly the Rotating Equipment Supervisor atthe Clark Refining & Marketing refineryin Hartford, Illinois and is a RegisteredProfessional Engineer in that state.


The Mystery of Cooling Tower Pump Noise

by Steve Schmitz, Bell & Gossett


HANDBOOK?

The problem of ”cavitating con-denser water pumps” withadequate NPSH available isnot uncommon. We have dis-

cussed this phenomenon with cool-ing tower manufacturers and othercentrifugal pump designer/manufac-turer members of the HydraulicInstitute. All agree that this conditionoccurs predominantly in cooling tow-er applications. We see on averageone or two such cases per year. Atpresent, several theories have beenoffered to explain the cavitation-likenoise. None has been validated. Wedo know, however, the following:

1. The noise is very similar, if notidentical, to classical cavitation(resembling the sound of marblesbeing pumped).

2. The phenomenon can occur witheither a forced draft or induceddraft cooling tower.

3. The noise tends to be moreprevalent on negative suctionpressure systems, but it willoccur on positive suction pres-sure systems as well.

4. The introduction of small amo-unts of air to the pump suctionoften quiets the noise. Thisentrained air has little effect on apump’s life expectancy.

5. Such small amounts of entrainedair have little deleterious effecton other system components.However, each system must beanalyzed for possible harmfuleffects.

6. Unlike classic cavitation, throt-tling the pump discharge to alower capacity usually has littleimpact on the noise level.

In an effort to determine theprobable causes, we visited a siteexperiencing such complaints. We

conducted a detailed inspection andmade an audio recording of thenoise spectra for laboratory study.The analysis revealed the following:

1. There were no distinct frequen-cies.

2. The predominate noise measuredwas broadband, occurring above300 Hz.

Pump noise can have both liquidand mechanical causes. Bothsources produce acoustic pressurefluctuations that can be transmittedas audible noise.

For centrifugal pumps, mechani-cal noise is generally the result ofcomponent imbalance (impellerand/or coupler), coupler misalign-ment, components rubbing againsteach other, or improper installationof the base plate and/or motor.These problems generate distinctfrequencies equal to rotationalspeed and/or its multiples (1,2,3).Because the noise spectra did notreveal distinct frequencies, wedetermined that this noise was not

mechanically generated.Liquid noise is produced directly

by water movement and is fluiddynamic in nature. Turbulence, flowseparation (vortex), cavitation,water hammer, flashing andimpeller interaction with the volutecutwater are all examples of fluiddynamic noise sources.

According to the Pump Handbook,2nd Edition, by Igor J. Karassik,there are generally four types ofpulsation sources in pumps that arethe result of liquid noise:

1. discrete frequency componentsgenerated by the pump impeller

2. broadband turbulent energyresulting from high flow velocities

3. impact noise consisting of inter-mittent bursts of broadbandnoise caused by cavitation, flash-ing and water hammer

4. flow-induced pulsations causedby periodic vortex formationwhen fluid moves past obstruc-tions and side branches in thepiping system

Figure 1. Incorrect suction piping and reducers installed upside down can makeit possible for air to collect, creating an obstruction.


Discrete frequency, item 1, canbe ruled out in this situation. Aspreviously mentioned, we did notfind distinct frequencies such as thevane passage frequency and/or itsmultiplies. The frequencies wouldhave been present if an interactionhad occurred between the impellerand volute cutwater.

Items 2, 3 and 4 are generallyidentified as broadband noise andwould occur in the 300 Hz andabove frequency range as identifiedon our noise spectra. Therefore, webelieve the noise was being generat-ed by one or more of these liquidsources.

The pump noise we heard waslike that of cavitation. Pump cavita-tion results from the formation ofvapor bubbles when the localizedstatic pressure is lower than thevapor pressure of the liquid beingpumped. To evaluate a pump forclassic cavitation (NPSHR greaterthan NPSHA), close the dischargevalve, thus pushing the pump backon its curve toward shutoff. Thenoise should diminish significantlyif it is originating from classic cavi-tation because lower pump flowsrequire reduced NPSHR. If thenoise continues, the cause is likelyto be entrained air.

We determined that classic cavi-tation was not occurring, as theoperating suction pressure mea-sured 30 feet above vapor pressure.Thus, the NPSHA was approximate-ly twice that required by the pump.For that reason, we knew the pumpwas not cavitating because of insuf-

ficient NPSHA. If not classic cavita-tion, then, what was causing thenoise?

It is a well-documented fact thathighly aerated cooling tower watercan contain as much as 4-6% excessair. This excess air increases thepotential for a noisy pumping instal-lation. The excess air absorbed inthe cooling tower comes out of solu-tion as it flows through the pipingand becomes entrained air. Suctionvelocities are often high enough topull the air through. However, airsometimes collects in an area of thesuction piping or the impeller eyeitself, creating an obstruction(Figure 1). As the liquid passesthrough this restricted area, itsvelocity increases, creating an areaof reduced pressure. At this point ofreduced localized pressure, watervaporizes, with the resulting bubblespassing into the pump impellerwhere, as the pressure increases,they collapse and produce ”cavita-tion.”

Several noise control techniqueshave been successfully employed inthe past to mitigate excessive noise.They include:

1. Increasing or decreasing thepump speed to avoid resonancesin the mechanical or liquid sys-tems and/or reduce the pumpNPSHR.

2. Increasing liquid pressures(NPSHA, etc.) to avoid cavitationor flashing; decrease suction lift.This could include raising thetower, lowering the pump orstraightening the suction piping

(see “Other Contributing Factors”)to reduce friction losses.

3. Modifying the pump so that theclearance between the impellerblade tips and casing cutwater(tongue) or diffuser vanes isincreased.

4. Injecting a small quantity of airinto the suction of a centrifugalpump to reduce cavitation noisesby providing a shock absorbingcushion that minimizes theimpact of recondensation ofwater vapor within the pump’simpeller.

Injection of small amounts of aircan usually be accomplished quick-ly and easily in the field with mini-mal expense. Small amounts ofentrained air usually cause no prob-lem in the cooling tower/condensercircuit. Bell & Gossett thereforeconsiders this alternative as desir-able and recommends its applica-tion as a solution to many fieldproblems or, at a minimum, as ananalytical tool.

Other Contributing FactorsIn addition to the techniques out-

lined above to reduce or eliminatenoise, attention must also be givento two other factors that can exacer-bate the situation: vortexing of theliquid in the tower pan, which is themost common source of air within apump (Figure 2), and the suctionpiping arrangement (Figure 3).

Vortexing of Liquid in the Tower PanThe amount of entrained air

caused by vortexing depends on sev-

EccentricReducer

5 to 10 pipediameters

Straight PipeFigure 2. Cooling tower vortexing is themost common source of air in the pump Figure 3. Proper suction piping


eral variables, but particularly thevortex size and the submergencelevel of the pump suction pipebelow the water level of the pan.The most common method of elimi-nating vortexing in the tower pan isby including baffle assemblies thatprevent vortexes from forming.Raising the fluid level in the pan to asufficient depth can also solve thisproblem.

Suction PipingCoupled with the vortexing phe-

nomenon, or by itself, improper lay-out of the pump suction piping canbe a significant contributor to thegeneration of pump noise.

Friction losses caused by under-sized suction piping can increase thefluid’s velocity into the pump. Asrecommended by the HydraulicInstitute, Standard ANSI/HI 1.1-1.5-1994, suction pipe velocity shouldnot exceed the velocity in the pump

suction nozzle. In some situationspipe velocity may need to be furtherreduced to satisfy pump NPSHrequirements and to control suctionline losses. Pipe friction can bereduced by using pipes that are oneto two sizes larger than the pumpsuction nozzle in order to maintainpipe velocities in the 5 to 10 ft/srange.

Eccentric reducers used to stepdown to the pump flange from thelarger suction piping can also be aculprit if they are used improperly.At the problem facility discussedearlier, the reducer was installedupside down, with the flat side onthe bottom (Figure 1). If the liquidcontains air (or vapor), as it did inthis case, the air can becometrapped in the sloped area of thereducer now located on ”top.” At aminimum this will obstruct the flowpassage, causing higher velocitiesand thus localized vaporization. Iftransported into the impeller, thetrapped air can create a momentarychoking that could even result inshaft breakage.

Elbows used on the pump suctionflange, while convenient, can causean uneven flow of liquid into theimpeller if the elbow bend is alongthe axis of the pump shaft (Figure 4).If the elbow is a short radius design,its use may unintentionally createturbulence that produces entrain-ment that can, and does, worsennoise problems. The addition of asecond elbow only increases theproblem, especially if the elbow hasbeen placed at a right angle to theexisting elbow.

Numerous technical publica-tions, as well as the HydraulicInstitute itself, state that systems

should have a minimum of fivepipe diameters of straight run ofpipe before the pump suctionflange to allow for a smooth unim-peded flow to the impeller (Figure3).

System strainers need to be locat-ed on the discharge side of thetower pumps and not on the suctionside (Figure 5). On another project,the location of basket strainersdirectly in front of the suctionflange on a large HSC pump result-ed in an unexpectedly high pressuredrop. This contributed to poorpump performance in that installa-tion, as well as higher pump noiselevels.

ConclusionIt must be understood that each

job site has its particular set of oper-ational requirements and, therefore,there is no single solution to themystery of cooling tower noise. TheHydraulic Institute has published astandard that provides recommen-dations for an ample margin of safe-ty between NPSHA and the pumpmanufacturer’s published NPSHR.The safety margin would be a mini-mum of 1.7 times NPSHR orNPSHR plus 5 feet, whichever ishigher. ■

Steve Schmitz is Senior ProductLine Manager, Centrifugal Pumps &Engineered Products, at Bell &Gossett. He has been with the compa-ny for 14 years and has written manyarticles on pump application and oper-ation. He is the concept originator andwas a member of the developmentteam for the successful ESP softwareprogram. He is also a member of theHydraulic Institute.

RecommendedInstallation

Long RadiusElbow

NotRecommended

Inlet Parallel toPump Shaft

Figure 4. Correct and incorrect elbowinstallation configurations

Strainer

Triple DutyValve

Service Valve

Concentric ReducerLong Radius

Elbow

Figure 5. Strainers need to be located on the discharge side of the tower pumpsand not on the suction side.


There was no mistaking thesound—that rushing metallicthud. As the valve slammedshut, a pressure wave travel-

ing at 4000 feet per second andmore than five times normal systempressure instantly crashed into theone-way check valve protecting themain supply pump from back flow.Once again, the integrity of the sys-tem held, but everyone knew it wasjust a matter of time.

This article addresses the all toocommon scenario of working nearpumping systems that are on theedge of catastrophe. The causes,results and solutions to the problemof surge or “water hammer” will bediscussed, as will the different butrelated problem of pump-inducedpulsation. To look at the practicalside of surge and pulsation, we willtake a trip through an imaginary,but realistically depicted, modernmanufacturing plant.

To set the stage for our tour, weneed to describe hydraulic condi-tions that set up the potential forthese problems to occur. Both surgeand pulsation in a liquid handlingsystem are the result of uncon-trolled pressure waves caused by anabrupt change in flow, either direc-tional or volumetric. Liquids con-tained in an enclosed system (pip-ing) have a physical volume; there-fore, a mass can be measuredand/or calculated. We can thendetermine the acceleration forcesneeded to move that given mass.Once in motion, the mass will stayin motion as long as enough force is

applied to overcome friction lossplus any gravitational component. Itcan be said that hydraulic equilibri-um is reached when the fluid isflowing in a laminar state.

Since for all practical purposesliquids are not compressible, forceor energy is not absorbed into thefluid but rather is transferredthrough it. The kinetic energy of themoving fluid will exert all the forceit has acquired to resist any condi-tion that tries to cause a change inits velocity. Depending upon themass and velocity of the liquid andthe rate of change applied to it, verydestructive forces can be generated,leading to catastrophic componentor system failure.

The critical consideration for thisdiscussion becomes the rate ofchange in energy for any givenmass. For example, consider theanalogy of a battleship moving atfull speed. If the engines arestopped, the ship will travel fivemiles in bleeding off the kineticenergy through water friction beforecoming to a stop. No damage wouldbe incurred because the massiveamount of energy involved isallowed to change form slowly.However, if that same ship at fullspeed were to run squarely into anaircraft carrier, all the kinetic energywould be concentrated and wouldchange form very rapidly—in a mat-ter of seconds. Since the greatermass of the carrier could not absorbthe kinetic energy as fast as it wasbeing delivered, the energy wouldbecome concentrated at the point of

impact.In liquid transfer systems, kinetic

energy is generally observed aspressure. Therefore when fluidvelocity is changed, the result is anincrease or decrease in pressure.Commonly, changes in velocityoccur when pumps are started orstopped (either intentionally or dueto failure), fluid flow directionchanges abruptly, pipe diameterschange abruptly or a quick-closingvalve shuts.

SurgeAt least in potential, a quick-clos-

ing valve (generally one that closesin less than 1.5 seconds) representsthe most dangerous condition. Sinceliquid efficiently transfers energyrather than absorbs it, a sonic orpressure wave is created. The inten-sity of the pressure shock wave isdirectly proportional to the speed offlow before the change in velocityand to the speed of propagation ofthe sonic wave created.Unrestricted, the pressure spike inthe liquid will rapidly reach thespeed of sound moving through liq-uid (approximately 4700 feet persecond). This is more than fourtimes the speed of sound throughair. It is mathematically possible tocalculate this pressure increase. Fordesign purposes, the increase inpressure can be determined by therule-of-thumb formula:

P = 60xVSWhere P = The increase in pres-

sure over the steady-state systemflowing pressure

V = Flow in feet per second of


HANDBOOK?Controlling Surge andPulsation Problems

The plant may be imaginary, but the problems are real. Take the tour and learn the mostcommon causes of problems like water hammer, and how you can prevent them.

by Gary Cornell, Blacoh Fluid Control, Inc.

the fluid before valve closureS = Specific gravityFor example, consider a 2” pipe

carrying 60 gallons per minute, witha system pressure of 100 psi and aspecific gravity of 1.2, and a valveclosing in one second. The 2” pipeand gpm equates to a flow rate of 6feet per second. In this scenario, thepressure increase above normalflowing pressure would be 432 psi,creating a total spike pressure of 532psi.

If the valve were shut in half asecond, the peak pressure woulddouble to 1064 psi. Conversely, ifthe valve were to close in 2 seconds,the peak total pressure would behalf of 532 psi, or 266 psi.

Note that the pressure increase isindependent of, but cumulative to,the normal system pressure. Forexample, a system pressure of 50psi flowing at 10 feet per secondwould have the same pressure risefor a given rate of change in veloci-ty as a system pressure of 300 psiflowing at 10 feet per second. Theproportional pressure change, how-ever, would be a much higher per-centage. Therefore, regardless ofwhether it is low or high pressure,the system may not be able to han-dle the pressure wave.

Clearly, when a mass such as liq-uid in a pipe is in motion and itsvelocity is changed, there is thepotential for a catastrophe.

PulsationUnlike surge, pulsation is the

rapid uncontrolled acceleration anddeceleration of units of energy. Inthe context of this article, theseunits of uncontrolled energy areactually slugs of liquid movingthrough a pipe. The degree of pulsa-tion in them is usually designatedby frequency in Hertz and a pres-sure amplitude (See Figure 1).

Outwardly, pulsation is usuallyobserved as component vibration orrapid gauge fluctuation. On an oscil-loscope it appears as a sinusoidalcurve (waves of peaks and valleys).Pulsation can occur and/or be influ-enced by the specific harmonics ofvarious components in a liquidtransfer system. Piping, valves andmechanical movement—and systemdesign itself, to a certain degree—

combine to influence measurablevibration.

However, the system componentthat instigates the pulse generationis the pump—specifically, a recipro-cating, positive displacement pump.This type of pump creates itsmotive force by repeatedly captur-ing and expelling a predeterminedslug or volume of liquid. It does thisby using inlet and outlet valves,which account for rapid accelera-tion and deceleration of fluid.Pulsation, then, is a rapidly repeat-ing change in energy form.Depending upon its frequency andamplitude, the potential for cata-strophic system component failureis very real. A simplistic comparisonwould be to the weakening effect ofthe human arterial system caused bya constant elevation in heart rateand blood pressure.

The major pulse in a pumpingsystem will be at the frequency ofthe plunger or piston speed timesany multiplicity factor. By way ofreference, a reciprocating pump’spulsation is generally described ashigh amplitude, but low frequency,as compared to the high frequencybut low amplitude frequency of acentrifugal pump’s impeller vane.Seldom are liquid handling systemsdesigned with an analysis of thetotal system harmonicsthat will occur.Whenever a recipro-cating pump is calledfor, at the very leastconsideration must begiven to the potentialeffect of the pulsationgenerated. In mostcases, minimizingpump-generated pulsa-tions will provide suf-ficient system protec-tion.

However, if theneed is to eliminatetotal system vibration,rather than pump-gen-erated pulsations, thenboth system and pumpharmonics must beconsidered. In suchcases, if there happensto be a harmonicmatch of frequencies,pressure amplification

can be many multiples of that gen-erated by the pump alone. This ismore likely to occur at higher fre-quencies such as those generated bya centrifugal or multipiston pump.

The TourNow that we have explained fluid

movement and discussed the unde-sirable effects that changing dynam-ics can produce, it is time to visitour imaginary plant (Figure 2). Forour tour we’ve selected the QualityPaint Company. QPC is one of thelargest paint manufacturers in theUnited States, with eight plants fromcoast to coast. They specialize inwater soluble paints using state ofthe art manufacturing equipment.

Stop 1: Tank Farm TransferPumps

As we pass the guard shack, thefirst thing we notice is a tank farm.

QPC purchases titanium dioxide(TiO2) solution in bulk and stores itin eight 10,000-gallon tanks. TheTiO2 is transferred into the facilityas demand requires. For several rea-sons, including the abrasive natureof the material and the sheer sensi-tivity and operational characteristicsrequired, the transfer pumps speci-fied are air operated doublediaphragm units (AODD). Each tank

Pulsations of ± 30 psi have been reduced to ± 3 psi witha SENTRY III Dampener

Figure 1. Undampened vs. dampened pressure pulsa-tions in a metering pump



has its own pump supplying prod-uct to a common header pipe goingoverhead into the plant. Eventhough AODD pumps are the bestchoice for this application, theyhave a rather poor tolerance forhigh inlet pressure. When the inletpressure approaches 12 to 16 psi,pump diaphragm life starts to deteri-orate rapidly. Since QPC’s TiO2tanks are 30 feet tall and their partic-ular mixture has a specific gravity of1.3, the static inlet pressure is 16.8psi. (feet in height x .432 x 1.3 s.g.)

AODD pumps are positive dis-placement, so they have inlet valveballs that act exactly like quick-clos-ing valves. From our previous dis-cussion we know that quick-closingvalves create instantaneous pressurespikes several times higher thanflowing pressure. Since 16.8 psi isabove the pump manufacturer’smaximum allowable inlet pressure,we can predict that there will beproblems. What happens in thepump is pretty straightforward. Thespike created when the inlet valveballs close on the liquid chamberseeks to move to a lower pressurearea. Since the inlet valve on thepump’s other chamber is simultane-ously opening and the motivediaphragm is creating a vacuum,this becomes the low pressure area.The pressure spike rushes in andslams against the diaphragm. Thisdistorts and stresses the diaphragm,leading to premature failure.

PositioningWe see that the pump is located

at the bottom of and just two feet or

so away from the tank. Because it isthis close to the tank, accelerationhead is not a significant factor. Onemethod of reducing high inlet pres-sure is to reduce the height of thetank, but this is not practical here.The AODD pump has several fea-tures that make it the best choice forthis application. So what can whatcan be done? A practical and economical solution is to install aninlet stabilizer as close to the pump’sinlet as possible, but within 10 pipediameters.

Inlet FactorsAn inlet stabilizer is a hydro-

pneumatic device consisting of apressure vessel with an elastomericbladder or diaphragm inside it thatseparates a compressed gas chargefrom the liquid being pumped. Theinlet stabilizer acts literally like ashock absorber and receives thepressure spike created when thepump’s inlet valve closes. An inletstabilizer is typically precharged to50% of the static inlet pressure.When properly charged and sized, itwill minimize the stress on thepump’s diaphragms by temporarilyaccumulating liquid and absorbingthe pressure spike. This minimizesstress on the pump’s diaphragms.

Discharge FactorsNow that the pump has been pro-

tected, the discharge needs to beexamined. The AODD pump pro-duces a pulsating flow. As previous-ly mentioned, pulsation is the rapidacceleration and deceleration of liq-uid caused by reciprocating action

of the positive displacement pumpin conjunction with the quick open-ing and closing of the pump’s dis-charge valves. This pulsing flow willbe observed as pipe vibration asenergy rapidly changes form.

Vibration Concerns—Mechanicaland Liquid

There is also mechanical vibra-tion caused by the shifting of pumpcomponents. And remember, if bothmechanical and hydraulic pulsescoincide, vibration will be increasedby a multiple factor. Due to theoverhead piping configuration,vibration is a serious problem.(Photo 1)

There are several ways to mini-mize the hazards caused by thisvibration. Larger diameter pipe canbe used to reduce back pressure.Thick-wall pipe can be installed todelay the eventual pipe fatigue. Andpipe braces (not hangers), used forsupport, can absorb some of thevibration. In effect, the system canbe overbuilt to absorb and dispersethe vibration caused by pulsation.Back pressure valves and orificescan provide some dampening effectbut usually at the expense of effi-ciency. All of these solutions, how-ever, simply address only the symp-toms of pump vibration and pulsa-tion. They will be costly and onlydelay inevitable component and/orsystem failure.

The most economical and properway to minimize a pump’s mechani-cal vibration is to use some type offlex coupling between the pumpand the discharge pipe. This isolatesthe system and prevents componentdamage. Several companies supplythis product. Sometimes, however,all that is required is a length of

Photo 1. Transfer pumps fitted withpulsation dampeners to alleviateshaking of overhead piping.

Figure 2. The imaginary Quality Paint Company plant


hose reinforced to withstand systempressure. As a rule of thumb, thehose should be at least 15 pipediameters long.

The most economical and effi-cient way to dampen the liquid’shydraulic pulses is probably a pulsa-tion dampener. This device is simi-lar in construction to the inlet stabi-lizer, but it must be pressurized to80% of the liquid flowing pressure.The pulsation dampener will absorbthe spike created by the rapid accel-eration of the liquid during the dis-charge stroke of the pump. At thesame time it will accumulate asmall amount of the liquid. Whenthe pump shifts, pressure is momen-tarily reduced as discharge pressureis lost and the dampener releasesaccumulated fluid back into thepipeline, filling the void created dur-ing the pump shift. This will mini-mize vibration. Also, liquid flowdownstream of the dampener willnow be in a continuous steady staterather than a pulsating start/stopmode.

Stop 2: From Mixing Tanksto Blenders

As we continue our tour andenter the plant, we see large mixingtanks where the paint is blended.Since the main ingredient in waterbased paint is water, we see an 8”water line running overhead and thelength of the plant. Coming off thismain header are six 4” branch lines,one to each blending tank.Following the main line back, weobserve that the water flow is pro-duced from a holding tank by an 8”end suction centrifugal pump pro-ducing 1500 gpm at 180 feet ofhead. A one-way stop check valveprevents system back flow, protectingthe centrifugal pump and creating apressure tight seal. The centrifugalpump starts on demand. The demandis based upon pressure changes in thesystem when valves at the blendingtanks open. These valves must be thequick-closing type because when apredetermined weight of water is letinto the blending tanks, flow mustimmediately stop.

From our earlier discussion onsurge or ”water hammer,” we knowthat as a mass changes velocity,

there is the potential for rapid ener-gy transformation. In this waterfeed system, we are faced with thefollowing potential problems:1.Rapid start-up of the centrifugal

pump against a pipeline full ofstatic fluid.

2.Rapid shutdown of the centrifugalpump by operator or by motorfailure.

3.Rapid closing of the valves at theblending tanks.Although there can be other

design factors of the system thateither minimize or exacerbate suchhydraulic problems, the items aboverepresent the greatest potential fordisaster.

Rapid Pump StartQPC’s centrifugal pump will start

automatically when the systempressure drops due to the openingof a blending tank’s valve. Whenthis occurs, the pump will throwwater into the 8” pipeline that isfilled with a stationary water col-umn. As the rapidly moving watercollides with the stationary column,a pressure (energy) spike will occur.This is similar to an automobile run-ning into a block wall. Great stresswill be put on the system. To pre-vent this from happening, severaloptions can be considered:

Slow or soft start pump motors willintroduce water into the systemslowly and minimize the pressurespike. This is a costly but effectivesolution.

Slow opening isolation valves willbe closed when the pump starts up.They will open slowly, allowinggradual pressure and flow increasesto equalize force in the stationarycolumn of water. This solution canbe complicated and expensive froma control point of view, but it willminimize the pressure spike.

Surge tanks are closed vesselswith air trapped in them. When thepump is started, initial water flowwill enter the surge tank until sys-tem pressure is equalized. The surgetank’s major drawback is that the airtrapped in it will be absorbed, creat-ing aerated fluid that is undesirablein paint products. In addition, oncea waterlogged condition exists, allcushioning effect will be lost.

Surge suppressors are similar in

design to surge tanks, but there isan elastomeric bladder inside sepa-rating the fluid from a compressedgas charge. By capturing the gascharge, the proper pressurerequired for the specific applicationcan be maintained. The suppressorshould be charged to approximately85% of normal operating pressure.A surge suppressor is an effectiveand cost-efficient way to minimizethe pressure spike that occurs dur-ing pump start-up.

Rapid Pump ShutdownWhen the pump is turned off, the

flow at the pump’s discharge stopsquickly. This creates a low pressurearea. Fluid column separation canoccur as momentum temporarilycontinues to move the mass ofwater down the pipeline. As soon asfriction acts to slow the forwardmotion of the liquid column, it willreverse and travel back toward thelow pressure area at the pump dis-charge. The reversal will travel atthe same speed as its initial forwardmotion. Depending upon the initialflow rate, the pipe gradient and thefluid mass, the pump casing will bestressed, with a potential for failure.Also, pump seal integrity can belost, and the impeller can becomewarped. Additionally, systems with

Photo 2. Metering pump installed witha pulsation dampener to inject chemi-cals into a process line.


a check valve can create a waterhammer effect with catastrophictransient pressure spikes when thereverse flow hits the one-way checkvalve. There are several solutions tothis problem:

Controlled-close check valves canbe used to time the closing period.A dashpot device or motor controlon the valve can be used to accom-plish this. If we control the rate ofchange in the velocity of flow, wecan minimize the amplitude of thepressure spike. This type of controlvalve works well in conjunctionwith check valves, but they must bemaintained for reliability, and theycan be an expensive option.

Surge suppressors can be installedjust downstream of the check valveor, if no check valve is used, at the

pump discharge. Properly sized andprecharged to 50% of operatingpressure, a surge suppressor willaccomplish two things. First, being50% precharged, the vessel willaccumulate fluid during pump oper-ation, and this accumulation will bereleased at shutdown to prevent col-umn separation. Second, the sup-pressor will absorb the pressurespike generated as the water col-umn reverses against the checkvalve or pump. The suppressor is astraightforward and economicalapproach.

Pump motor failure or loss ofpower is another contingency thatmust be considered. While modernelectric pump motors are extremelyreliable, failures do occur. Poweroutages and surges must also betaken into account. When the motorstops, the same condition exists asat pump shutdown. That is to say,column separation can occur andflow will reverse against the checkvalve. The worst possible situationis where power is only momentarilyinterrupted and then, just as theflow column is returning to thepump, the pump restarts. Thiswould be similar to a head-on colli-sion of two automobiles traveling athigh speed.

Options to control these situa-tions are few. Certainly an interruptor time delay switch could beinstalled at the motor to prevent animmediate automatic restart.Controlled close and open checkvalves will not be an option becausethey probably would not be opera-ble or would not react fast enough.

The best choice here would be asurge suppressor. The compressibili-ty of the gas in the suppressor willreact instantly to the transient pres-sure spike.

Quick-Closing ValvesMoving down the water feed line,

we now need to determine whatwill occur when the valves at theblending tanks are closed quickly.For purposes of this discussion, wewill define a quick-closing valve asone that shuts in 1.5 seconds or less.It is important to remember, howev-er, that as valves get larger, they canstill cause problems even if theyclose more slowly.

But what actually occurs to cre-ate a potentially catastrophic spike?When one of the valves closesquickly, flow there is instantlystopped, but the column of waterbehind it will still be moving. Thinkof a speeding train with the engineabruptly hitting a stationary block-age. The engine stops, but the carscontinue moving forward. The onlydifference between a train andwater flow is that the train is notcontained, and the cars derail. If thepressure spike does not rupture thepipe or some other component, thecompression wave created willreverse and travel back down thepipe toward the pump at the speedof sound. When the wave hits thecheck valve or pump, it will againreverse and continue to reverberateuntil something finally breaks or theenergy dissipates due to frictionloss. Even if nothing fails initially,the system is under repeated stress,a situation that will ultimately leadto fatigue failure.

What can be done to preventthis? The pressure spike can beeither released or absorbed. The fol-lowing options are among the mostaccepted solutions:

1. Rupture discs 2. Pressure relief valves3. Slow closing, timed valves4. Surge suppressors

Rupture discs are simply plugs thatwill break at a lower pressure thanany other component in the system.They are usually a ”last resort” pro-tection and are not normally used tocontrol pressure spikes because theliquid released into the environmentis often hazardous, costly, or justdangerous in being released at sucha high pressure.

Pressure relief valves are valvesthat will open at a predeterminedpressure. As the pressure spikebuilds, a relief valve will open, andfluid will be released. Either a hold-ing tank or a return pipe systemmust be installed to exercise thisoption. Valve sizing and relief pressuresettings are critical to minimize a spikeas quickly as possible. This approachcan be costly, and relief valve reliabili-ty must be tested regularly.

Slow closing valves will solve the

Photo 3. Pulsation dampened air oper-ated diaphragm process pumps

Photo 4. Paint filling machine dispens-ing into five-gallon buckets. Accuracyand reduced splashing are achievedby the smooth flow provided by pulsa-tion dampeners.


problem if there truly is no need toclose the valve quickly. In QPC’sapplication, however, it is necessaryto stop flow instantly, so this solu-tion is not an option.

Surge suppressors provide a solu-tion that will absorb the pressurespike by momentarily accumulatingthe flow of liquid as the valve clos-es. Because of the speed of propaga-tion of the transient wave created,the surge suppressor must beinstalled directly upstream from thequick-closing valve and in no situa-tion further away than 10 pipediameters. Since the full capacity ofthe suppressor must be available toaccept the accumulation of liquidwhen the valve closes, the suppres-sor must be precharged to 95% to98% of system pressure. It must alsobe properly sized so it can providethe proper compressed gas cushionand momentarily accumulate a pre-determined amount of liquid. Asurge suppressor installed at eachquick-closing valve for this applica-tion will provide an economical andreliable protection.

In addition to the water supplysystem, we observe a bank ofmetering pumps that are used toinject precise amounts of fungicideinto the paint blenders (Photo 2).This is accomplished with small sin-gle-diaphragm metering pumps dis-pensing the chemical through a pre-cise mass flow meter. Because of thereciprocating nature of the meteringpump, we know the discharge flowwill pulsate. In many applicationsthis is not a problem, but becausethe flow meter cannot accuratelymeasure a pulsating flow, a pulsa-tion dampener must be installed atthe pump’s discharge. If properlysized and charged, the dampenermaintains the mean system pressurewithin 1%, which is sufficient forthe meter to function properly.

Stop 3: Packaging the Finished Product

From the blending tanks, AODDpumps transfer the paint to holdingtanks before it goes to filling equip-ment. The AODD pumps are all fitted with discharge pulsationdampeners to prevent rubber hoses,

which connect the pumps to thetanks, from ”jumping” around theplant floor and endangering employ-ees. Hose jumping results from pul-sations created by the start and stopaction of the pump. In addition, thejumping will wear holes in the rein-forced hose wall as it rubs on theplant floor. Since it installed damp-eners QPC has been able to reducehose replacement greatly.

As we pass the filling equipmentfor buckets and cans of paint, weobserve several more pulsationdampeners, again installed in con-junction with AODD pumps (Photo3 and 4). The filling machines can-not be accurate with pulsatingflows. The manufacturer of the fill-ing equipment installs dampenerson its equipment, so that whetherbuckets and cans are filled by flowmeters or by bulk weight measure,the flow is laminar and measurableby the instrumentation.

The FutureThe last stop on our plant tour is

the research and development area.QPC is committed to ongoing prod-uct development, and today techni-cians are testing a new proprietarypaint for spray applications. Theyare using a piston type spray pump,which produces a reciprocatingflow. Their goal is to produce apaint that will disperse well throughspray nozzles for coating purposes.They will need to use a pulsationdampener with this pump for tworeasons. First, system vibration atthe pump’s high cycle rate can leadto system fatigue and eventual fail-ure. Second, without the dampener,the spray pattern produced will bewavy and inconsistent as the sprayrises and falls in synchronizationwith the pressure pulse from thepump. The inconsistent pattern willrequire over-spray to coat the testboard. This results in paint wasteand a non-uniform coverage.Properly sized and charged, a pulsa-tion dampener can provide therequired level of dampening toallow for a smooth and continuousflow from the nozzles.

Although our tour is finished,there are many other pump systems

in a typical plant that would benefitfrom pulsation and surge control.Each should be analyzed withregard to potential hazards that mayresult from a change in velocity ofthe flowing fluid or from the pulsa-tions created by a reciprocatingpump. Examples include filter presssystems, where the pulsating flowfrom an AODD pump can damagefilter media or ”cake,” and in-linemixers, which benefit from a steadystream of injected fluid as opposedto slugs of liquids.

Our tour was not designed to be adetailed analysis of every system,but rather to highlight the kinds ofdamage that can occur in any systemthat uses pumps and valves. Nomanufacturing plant need subjectitself to the potential hazards of pul-sation and surge. I have made noattempt to explain sizing, materialsor other parameters involved inusing the preventive devicesreviewed. In some cases, complicat-ed mathematical formulas arerequired to make such determina-tions. There are several good manu-facturers of the various devices dis-cussed. You should consult withthese manufacturers or their distrib-utors to determine the proper prod-uct(s) for your specific application. ■

Gary Cornell, president of BlacohFluid Control (Riverside, CA), hasworked in the reciprocating pump indus-try for more than 20 years. He has aB.S. degree from California PolytechnicUniversity, and has had several articlespublished on the subject of pulsation andsurge control.

ReferencesKarassik, Igor, William Krutzsch

and Warren Fraser. Pump Handbook.McGraw Hill (1976)

Wachel, J.C. and S.M. Price.Understanding How PulsationDampeners Work.

The American Society ofMechanical Engineers PipelineEngineering Symposium (1988)

Young, Winston. The YoungEngineering Technical Manual.

Young Engineering, Monrovia, CA


Many pumping industrystart-up stories are tales ofhorror. Here are just a fewof the most common prob-

lems and some solutions that canovercome them.

Oil LeaksOil leaks are a typical start-up

nuisance that simply should nothappen. Most oil leaks occur at thegaskets, lip seals, pipe plugs, bear-ing isolators, oil bowls, breathers,oil lines and O-rings. Most of theseleaks can be avoided by using quali-ty parts and superior installationtechniques.

The invisible oil leak is a toughproblem to detect and solve. Onereason this type of leak occurs isthat an oil bowl or sight glass hasbeen installed on the wrong side ofthe bearing housing. The rotation ofthe bearings and/or the oil slingercauses the oil to be higher on oneside than the other. This gives afalse reading, and the lubricator

adds more oil than necessary. Thisexcess oil in the bearing housinghas two effects. The bearings willrun hotter than normal, and there isthe possibility of an external leak.

Oil leaks caused by vacuum inthe coupling guard can be a realbother and a headache to trou-bleshoot. These leaks occur whenthe coupling guard is mounted soclose to the bearing housing that thecentrifugal force of the couplingpulls air like a house fan.

Oil leaks in totally enclosed bear-ing housings usually result from anair pressure imbalance between theoil reservoir and the housing. Thisnormally occurs with the use ofcontact seals, such as magnetic orthe Inpro VBX positive vapor seals.Magnetic bearing seals are obvious-ly not vents for the housing, butneither are Inpro VBX bearing isola-tors. If an oil reservoir is used, itshould always be a closed systemlike the ones available from eitherTrico or Oil Rite. An expansionchamber or a directional vent valvemust be used on top of the pump toprevent moisture ingestion from

breathing. This arrangement willensure that as the air inside thepump expands and contracts, theair pressure between the oil reser-voir and the bearing housing staysin balance, thereby eliminating theleak.

It is common to have oil leaks inhorizontal pumps that, for spaceconsiderations, are stored in a verti-cal position. Contact seals preventthis type of leak. Always be sure touse good installation practices withany oil seal, taking care to avoid thesharp burrs that can make any goodseal look bad.

Start-up leaks on ANSI and otherpumps occur at the gasket betweenthe adapter frame and the bearinghousing. These leaks can be frombroken or folded gaskets or fromtrash holding the two faces open.However, this kind of leaking oftenresults from metal crowning at thethreaded area of the sealing surface.Crowning happens when thethreaded bolt holes do not haveenough of a bevel cut at the face,and the torque applied by the boltsforms a crown of metal around the

Avoiding Gremlins andAlligators at Pump Start-up

Is a checklist all you need? Or should you buy alligator repellent, too?

by Bob Matthews, GCI-Texas

Shaft

Cast IronHousing

Effected Area

�y

Figure 1a. Over-torque or tapped holeswithout chamfer will pull a crown out ofmetal on machined surfaces. Theseraised areas will cause many problems.

These raised areas can be indicated and removed with a mill file

Figure 1b. The raised areas can be easily identified and removed with a mill file.


HANDBOOK?


hole (Figure 1a). This problem iseasily resolved by adding a note tothe rebuild sheet to check for this.Filing with a large flat mill bastardfile or a carbon stone on the tappedsealing face will reveal the crowninstantly, and it can then beremoved (Figure 1b). RTV siliconecan be applied when raised areasexist, but this often puts the twomating faces out of square, initiatinga chain reaction of alignment prob-lems. I’ve mentioned only a fewproblems here. Be sure to use yourimagination and good troubleshoot-ing techniques.

Leaks and GasketsI have seen start-up leaks on split

case pumps stopped by the use ofthicker gaskets at the split lines.However, this causes the stuffingbox to become elliptical and differ-ent problems to occur (Figure 2).Technicians should begin by check-ing the manufacturer’s maintenancemanual for a specification on propergasket thickness. You can also deter-mine the proper thickness in thefield by assembling the two housinghalves using only half of the nuts.Once this has been done, you canmic the stuffing box at and perpen-dicular to the split line. Subtract theheight from the width to determinethe exact gasket thickness required.

Product LeaksProduct leaks are a giant head-

ache in centrifugal pumps. Manysuch leaks occur at the gaskets,flush lines, pipe plugs, O-rings,packing or mechanical seals. Themajority can be resolved by usinggood quality parts, superior installa-tion skills and technical adjustmentsduring start-up.

Leaks Caused by CracksProduct leaks in cast iron pump

cases and oil leaks in cast iron hous-ings result when cracks develop dueto over-torquing to compensate forgasket leaks. The most overlookedand easily detectable culprit is theimpact wrench being set on thehighest torque setting with anti-seize on the threads. To top it all off,the torque wrench is applied to thebolts after the impact wrench.Amazingly enough, the torquewrench clicks! Yes, the wrench mayclick to indicate it applied the

wrench setting, but more torqueactually exists because the impacthas surely applied too much torque.Important tip: If the torque wrenchdoes not move the fastener before itclicks, the torque is set too high. Evenif over-torqued bolts don’t causecracks that leak product at start-up,vibration can finish the work startedby the impact wrench, resulting inleaks. There is a great differencebetween the torque wrench settingsfor dry threads and those paintedwith anti-seize products. Be sure tochoose the proper torque for yourconditions. Failure to do so will like-ly result in your being bitten by anyor all of these start-up and extend-ed-run bugs. If you are not sure, youcan identify the torque applied bythe impact wrench by loosening thebolt with a needle type torquewrench. Next, make a visual inspec-tion of the bolt heads. To trou-bleshoot a historic problem of thistype, look for the mushroom edgemarks on one side of the flats madeby the impact wrench and fordepressions in the face of the cast-ing under the bolt heads.

LubricantContaminated lubricant can also

cause pumps to fail at start-up.Contamination can come in theform of water, steam, product,sweat, dust or dirt. Totally enclosingyour bearing housings is a good firststep to prevention. For example, inthe chemical and refining processindustry, most pumps are spared.One pump runs the majority of thetime, and another runs hardly at all.The spare pump with a long idletime can have start-up problemsdue to oxidation in the bearings.This is quite common in humidareas like south Texas, where I live.As the pump breathes, condensationforms on the walls of the bearinghousing. The result is rust. Aftertotally enclosing the housing, thenext step should be a switch to asynthetic oil.

Finally, an even better idea wouldbe to install an oil mist lubricationsystem for the bearings. Volumes ofdocumentation show that the costsaving benefits of this idea far out-weigh oil mist’s cost of installation.Some reasons are:•Less oil is used. •Air flows across the bearings,which helps to cool them.

•A small but constant supply ofclean oil is applied to the bear-ings.

•A slight positive pressure is heldwithin the housing to stop thedestructive breathing process. Oil misting will all but eliminate

the start-up bearing failures describ-ed here. And if you have totallyenclosed the housings, there won’tbe any housekeeping problems dueto stray mist.

The Break-in PeriodLack of technical maintenance

during the break-in period accountsfor most of packed equipment’sstart-up problems. At start-up andfor several hours afterward, packedstuffing boxes must have the atten-tion of qualified personnel. If pack-ing is not adjusted carefully,burnout can occur, causing produc-tion delays. Proper selection andinstallation techniques for packingare very important. Naturally, if you

Figure 2. The diameters of A and Bmust be equal. Gaskets which are toothick can cause problems. If runningpacking, the result is a leak. The quickindicator of this situation is that thegland will lock. If running mechanicalseals, the bolt circle will be out of lineand bind the seal’s internal parts.


use packing material that does notsuit the application or if you installit incorrectly, start-up problems areinevitable. It sounds simple, butmillions of dollars are wasted eachyear because packing is not selectedand adjusted properly. I have foundthat the Anti-Keystone design pack-ing is more consistent and easier tocontrol than others during start-ups,resulting in longer run times.

Mechanical SealsMechanical seals require start-up

attention but often go unchecked.Inspection at start-up is imperativeto ensuring the necessary flushrates. If you use a barrier fluid tank,someone must check for properoperation and effectiveness as thepump starts to run. Whether a sealflush or a barrier tank is used, quali-fied personnel must check the pip-ing against the specified require-ments to see that it is correct. Donot let anyone tell you it is all rightto put barrier fluid in a double ortandem seal other than from bottomto top. Clear-housed seal chambersshow that the trapped air does notleave the area during rotation but ispulled to the rotor, thereby starvingthe seal faces of lubrication andcooling. If there is a lack of availablequalified maintenance personnel,control the start-up electronically.

Dry RunningOnce every so often you may

hear of a pump started with nomedium, or the suction is dry. Asingle seal will burn up, and this isa guaranteed product leak.Sometimes the barrier fluid is leftout of the tank and the outboardseal of a double seal will fry. Thisturns the double seal into a singleseal and damages it beyond repair.A packed pump, which depends onproduct for cooling and lubrication,burns up without liquid in thepump at start-up. This glazes theinner packing surface, and a leakwill occur as operating pressure isreached. Level controls and othersignal devices will prevent many ofthese problems.

Dry ProductsDry products in pumps, products

with solids, and products that solidi-fy when left to dry out in the volutecan create several start-up prob-

lems:• The motor will pull high amps.• The shaft can break or bend.• Product will dry on the impeller,

causing imbalance and vibration.• Seal faces become glued together,

destroying them when horsepow-er is applied. Pumps that operate in such condi-

tions require the development of aflush procedure to eliminate thesefailures. To say the least, it is a goodidea to lock out the equipment andcheck for free rotation before start-up.

VibrationHere are just a few of the reasons

why vibration will kick your butt atstart-up. Balance specifications rec-ommended by some manufacturersand some engineers are not tightenough. The better you balance apiece of equipment, the better itruns. The ultimate compliment atechnician can get is for an operatoror supervisor to say that the equip-ment is not running when in fact itis! Again, I cannot adequatelyexpress how important it is to have acheck-off sheet for repairs. Do every-thing possible to make rotatingassemblies run smoothly. Put thatcoupling hub on the shaft when bal-ancing and take your tolerances past“just good enough” on the balancemachine.

Also, don’t overlook what I mightterm “flow imbalances.” When themass of a rotor is larger on one sidethan on the other, even if it is with-in dynamic balance specifications,the push creates a medium forceimbalance. Another example iswhere the impeller bore is off cen-ter and the vanes on one side arelonger than on the other. Thisimbalance transmits right to thebearings as a radial load and short-ens pump life.

LockupA pump can lock up at start-up

for many reasons. Lockups happenwhen bearings fail due to lack oflubrication. Leaks from neglectedand loose oil drain plugs can causethis problem. The preservativelubricant used for pump storagemust be compatible with the oilused during the pumping operation.Have your oil supplier’s engineercheck the preservative to be sure it

will blend with the oil in the bear-ing housing. Thread compound orTeflon tape will seal threaded plugsand pipe fittings from most leaks.When using these thread sealers, donot use too much because excessescan get into ports under bearingsand plug the housing’s oil returns.

Too small clearances in case wearrings can also cause lockups. Thishappens when the rings are setwrong, when temperature changescause expansion, when shaft runoutis excessive or when materials arechanged, such as going from carbonto steel to stainless. Understandingthe process in which the pump isrunning is important. Technicianswho lack this knowledge put them-selves at a great disadvantage. Goodtechnical information, skills andprocess knowledge are the keys toresolving these problems.

Unexplained LockupsSome bearing lockups at start-up

are seemingly inexplicable. My co-worker, Mike, told me of a bearinglocking up with adequate oil in thehousing. He immediately appliedtroubleshooting skills and found nocause. Mike changed from the firstoil used to Royal Purple Synfilm.This worked, but who knows why?Sometimes you must go with yourgut feeling. I greatly prefer to useengineered technology and experi-ence to solve problems. When Iasked for permission from RoyalPurple to use their name in this arti-cle, they gave me an explanation forthe phenomenon just described.“Synfilm” is the brand name usedfor their high film strength syntheticlubricant, which is derived fromtheir proprietary “Synerlic” additivesystem. The bottom line is thatSynerlic forms an ionic (electrostat-ic) bond with the metal surfaces,leaving a long lasting film on thebearing that is very good at start-ups. Reduction in the coefficient offriction when Mike added theSynfilm to the bearing housingenabled the machine to run.

Sleeve BearingsSleeve bearings in vertical medium-

lubricated pumps are a never-endingsource of problems. Poor shaft sup-port, dry starts, high friction and rapid


wear in abrasive services are some ofthe more common difficulties encoun-tered. One solution is the addition ofan enclosed tube around the shaftwith either a fresh water flush oroil/grease lubrication. This textbookfix has a high initial cost and a contin-uing expense in the supply of freshwater or oil/grease. It can also contam-inate the pumped medium. The solu-tion I am familiar with is the installa-tion of Thordon bearings. This compa-ny produces a range of medium-lubri-cated bearings offering low friction,dry start capabilities, good shaft sup-port and excellent resistance to abra-sion.

Running Backwards andFlowback

Running backwards at start-upwill spin some screw-on styleimpellers loose and jam them intothe volute, locking up the pump.Some impellers can be installedbackwards and still fit into thepump case and thus go unnoticed. Iknow this can happen because I waspart of this mistake once. What alearning experience! Pumps startedin reverse can have a variety ofeffects on a system, most of whichare very costly and dangerous.When install-ing a pump, alwaysverify that everything will rotate inthe correct direction.

Fluid back-flowing through apump at start-up will often breakthe pump shaft, and it can cause

any number of other problems.Backflows through leaking checkvalves or bypasses with a largeenough volume of medium will spinan im-peller backwards. Visualinspection of the pump before start-up is necessary to insure that thisproblem does not occur.

AlligatorsThese creatures have also been a

problem at start-up. I know youmust think I’m crazy, but this is atrue story. A young farmer went tostart his irrigation pump one dayand noticed that a foundation sup-port had been moved away from thedriver engine. Unknown to him, alarge alligator had moved into theneighborhood to feed on arearodents and had knocked out thesupport block. After repair andrealignment, the engine was readyto start. The next morning as myfarmer friend began to start hispump, he slipped on the morningdew, slid down to the suction pond,and came face to face with Mr.Gator. His escape was successful,and he will always remember thatparticular start-up. And I’ll bet youanything that this pump now has astart-up checklist!

Final TipsTo summarize, I will suggest that

you can improve start-ups in theseways:• Always have a hands-on mainte-

nance crew present to tune theequipment until the operation isrunning smoothly.

• Choose the best parts and lubri-cants to maintain long life whenrepairing pumps.

• Go that “extra step” with toler-ances. I don’t have engineeringdata to prove that tighter specsare better, but seeing longer life isproof enough for me.

• Use good pump check sheets.This is necessary, not optional.

• Make the rounds. Look the pumpover before start-up. Confidenceis important, and these pre-checks save money. (Is the flushwater on? Is the shaft going torotate in the right direction? Isthe oil level all right? Is the mistturned on? Is the pump case dry?Answer the applicable questionsfor your equipment.)

• Find out if application changeshave been made and if so whateffect they might have.

• Look out for that alligator! ■Bob Matthews is a Rotating

Equipment Consultant for GCI-Texasin Pasadena. He has more than 30years of experience in mechanicalapplications, 10 of which were hands-on pump repairs. He is a graduate ofLamar University. Bob is a frequentcontributor to Pumps and Systems. Ifyou have questions about the article,call him at (713) 473-7802 or send afax to (713) 473-4068.


Utility boiler feed pumps areamong the most criticalcomponents in a powerplant. They are used to pro-

vide feedwater to the boiler whereit is heated to produce steam. Thereliability of these large pumpsaffects the reliability of the entireplant because output must bereduced or stopped when thepumps are not able to supply waterto the boiler. The following caseanalyzes how Potomac ElectricPower Company (PEPCO) increasedits efficiency and reliability throughclose monitoring of its boiler feedpump operations.

Boiler feed pumps used in largepower generating stations must bedesigned to operate reliably in thesevere operating conditions of highheat, pressure and flow. A large sta-tion of 600 megawatts (MW) wouldtypically have two parallel pumps,each rated at about 15,000 hp,pumping feed water at about 500°F.The discharge pressure of suchpumps in a supercritical cycle isaround 5,000 psi, with a flow ofabout 5,500 gpm. Interestinglyenough, some utilities of compara-ble operating capacity are notequipped with two parallel pumps;rather they utilize a larger singlepump to provide their feedwater.Most of these pumps are variablespeed, and they are typically drivenby mechanical drive steam turbinesor electric motors through hydrauliccouplings. Recent technology hasintroduced variable-speed electricmotors, which have become analternative for driving these pumps.

PEPCO’s two Chalk Point Station

units,w h i c hwere com-pleted in 1964and 1965, had access to limitedsupercritical boiler feed pump expe-rience, as did its fellow utilities.Chalk Point’s primary pumps onUnits 1 and 2 were an early genera-tion supercritical design. Theseunits, like most early models of anyproduct, had design limitations thatdirectly affected their in-service per-formance.

Each unit generates 350megawatts and is equipped withtwo parallel boiler feed pumps. Thepump operating conditions areabout 2,800 gpm with a dischargepressure of 4,800 psi. Operatingtemperature is about 360°F. Thesepumps, although efficient, experi-enced complications early, withthree of the four pumps failingwithin one month of initial start-up.Over the years these pumps experi-enced low MTBF and high MTBR.Simultaneously, other utilities wereexperiencing a similar trend withthese early model pumps.

In 1979 a major boiler modifica-tion project was proposed for thisstation. The project would introduceadditional pressure drop in the boil-

er, thereby requiring additionalhead to be produced by the boilerfeed pumps. A study was done toreexamine the capabilities of theexisting boiler feed pumps. Theresearch indicated that performanceof the existing pump was marginalin providing the required flow at ahigher discharge pressure. Facedwith these findings and forecastingfuture demand for generation,PEPCO took an uncommon initia-tive in the utility industry and pro-posed the replacement of existingpumps with new pumps, primarilyto improve reliability.

Before purchasing the newpumps, a task force was convenedto determine the root causes of thepump failures and the correspond-ing requirements or system changesneeded to assure optimal efficiency.The managers at Chalk Point used aresearch project prepared by theElectric Power Research Institute(EPRI) for this purpose. Based on asurvey of a number of utilities,EPRI issued a report that correlateddesign features with reliability prob-lems. The institute identified a num-ber of potential problems, and thesewere compared to the Chalk Pointpumps. Specific problem areas iden-tified by EPRI that were common tothe Chalk Point pumps included thefollowing.

• Flat rise to shutoff of the pumphead/capacity curve

• High specific and suction specificspeed

• Marginal NPSH available/NPSHrequired

• Operation off Best Efficiency

Experience with Replacementof Boiler Feed Pumps for Reliability Enhancement

By Merwin W. Jones, Potomac Electric Power Company

Figure 1. Typical HDBpump


HANDBOOK?


Point (BEP)• Inadequate minimum flow

The utilities surveyed for thestudy, including Chalk Point, alsoexperienced difficulties with shaftseals, balancing mechanisms andinterstage takeoff seals. The pumpswere also suspected of having aproblem with inlet eye flow recircu-lation due to modified impellersinstalled several years earlier. Mostof these failures were related to themanufacturer’s design, particularly,close clearances used to enhancethe pump’s performance.

At the conclusion of the investi-gation, the task force recommendedreplacement of the old boiler feedpumps. PEPCO prepared a specifi-cation for new pumps thataddressed all the issues raised dur-ing the task force investigation.Although plant design created limitssuch as maximum shutoff pressureor suction pressure available to theboiler feed pumps, each deficiencyin the situation was examined care-fully and was specified as accurate-ly as possible. For example, thefeedwater system was given a com-plete test at minimum through max-imum flow to characterize the rangeof operating conditions. Smalladjustments were made in the testresults to account for proposed boil-er modifications.

Specific issues addressed in thespecification and a discussion of thefindings and approach used to elimi-nate them as a source of unreliabili-ty are given below.

Constantly Rising Head CapacityCurve from Runout Condition toShutoff Condition

This is a standard requirement inpump specifications. However, theshape of the curve is also important.The existing pumps had a sharp risefrom runout flow to about 75%flow, and a flatter rise at lowerflows. As a result, it was found thata change in flow from about one-half flow to shutoff resulted in anincrease in head of only about 25psi. This indicated potential instabil-ity at low loads. Fortunately, therewas a slight margin in the pipe’sallowable pressure rating—a circum-stance that enabled the pump

designer to increase the maximumallowable shutoff head slightly. Thiswas not considered a major risk,because we were still within therange of stresses permitted by thePower Piping Code, and becausethese pumps are turbine driven andreduce head at lower flows byreducing the turbine speed.

Specific Speed and Suction SpecificSpeed

The existing pumps were foundto have a high specific speed and avery high suction specific speed. Inthe EPRI surveys, pumps with highspecific speeds were found to have amuch greater failure rate. Althoughlimits were not placed on theseparameters in the specification, theywere considered in the evaluation ofnew pumps. Suction specific speedwas reduced from about 12,000 forthe old pumps to 8,700 for the newpumps, largely attributable to a dou-ble suction first stage impeller.Specific speed was also reduced dueto one additional stage in the newpumps.

NPSH Available versus NPSHRequired

NPSHA was marginal for theapplication, at least at the highestflows, so a great deal of attentionwas focused on this area. The sys-tem test was used to establish anaccurate NPSHA curve. Our investi-gation of new pumps showed thatthe criteria for establishing NPSHrequirements varied from vendor tovendor. NPSH is often based on theamount of cavitation to reduce thedischarge head by 1% or 3%. Forthis application the pump vendorwas required to provide NPSHrequirements based on a 0% headloss, and to perform tests at four dif-ferent flows to confirm the NPSHcurve proposed. To establish a 0%head loss NPSH level, the pumpswere tested at the vendor’s testfacility, and suction pressure wasreduced until it affected the totalhead produced by the pump. Datapoints were established to show theNPSH at about 1% and 3% headloss and then extrapolated to 0%loss. To allow for system transientsand upsets, the required NPSH wasspecified so as not to exceed 2/3 of

the available NPSH.

Operation Off Best Efficiency PointOperation at less than the design

Best Efficiency Point (BEP) is a com-mon cause of premature failures oflarge feed pumps. The usual expla-nation for this condition is that dur-ing the design process margins areadded for flow, piping pressure loss-es and other uncertainties in theapplication. The combined effect ofthese design margins results in apump that is sometimes substantial-ly oversized and therefore alwaysoperates below its design point.This concern was eliminated byusing the results of the feedwatersystem test. The test established theexact system head resistance curve,so the replacement pumps could bedesigned to operate near the BEP.

Inadequate Minimum FlowAlthough the existing recircula-

tion system had adequate minimumflow to prevent flashing of the feed-water, the criteria in recent yearshad been changed to provide for aflow at which the first stageimpeller operates in a stable flowregime. In the case of the selectedpumps, the minimum flowincreased from 550 gpm to 1100gpm. This necessitated changes inthe recirculation valves and piping.

Balancing MechanismThe thrust loads on a multi-stage

pump with a discharge pressure ofabout 4800 psi must be balanced bysome force within the pump.Although small pumps generally usea thrust bearing to absorb theseforces, a device of this kind is con-sidered impractical for large boilerfeed pumps. Most vendors havedeveloped hydraulic mechanisms tobalance the thrust.

The existing pumps at ChalkPoint were equipped with a balanc-ing disk. This design relies on asmall clearance to throttle the dis-charge flow into a balance chamber.Pressure changes open or close tomaintain the location of the shaftwith in a few thousandths of aninch. This mechanism was animportant issue at Chalk Point. Itwas our opinion that the change inoperation of this unit from full load


during the day to about one-thirdload at night increased the rate offailure of this component. PEPCOhas several pumps at other plantswhich utilize opposed impellers tobalance the thrust. By aligning halfof the impellers in one direction andthe other impellers in the oppositedirection, only minimal hydraulicbalancing is required. PEPCO’sexperience with these pumps insimilar cycling duties was consid-ered very good.

The company also conductedmany interviews with other utilitiesusing similar size pumps. It wasfound that units which cycled inload had higher rates of failure withdisk or balancing drum type thrustbalancing than those in similar ser-vice with opposed impeller designs.Our evaluation considered this fac-tor for new pumps.

TestingThe pumps were each given a full

flow and pressure performance test

at the vendor’s shop. The tests veri-fied performance in terms of flow,pressure, NPSH and efficiency. Thevendor was required to subject thereplacement pumps to additionaltests as well. These were proposedas a result of recurring failures tothe old pumps, as well as inresponse to claims made by thepump vendor.

For example, an interesting testwas done to demonstrate the abilityof the pumps’ shaft seals to with-stand loss of injection watermomentarily. Loss of injectionwater on a pump of this size usuallyresults in severe damage in only afew seconds. During the test, thepump was set to operate with mini-mum NPSH, and the injection waterwas shut off for one minute. Werequired that the pump show noincrease in vibration during thisperiod, and the shaft seals were tobe disassembled after the test andexamined for any signs of rubbing

or wear. These pumps passed thistest.

A particularly notable event dur-ing the NPSH tests occurred atabout 80% flow. The pump suctionwas throttled to reduce the pressureto 5.3 psig. In this condition thepump was passing 2,125 gpm at adischarge pressure of 4,200 psi. Thepump remained quiet with no out-ward signs of cavitation.

NPSH “Meter” During the system tests it was

found that hotter than normal waterwas sometimes entering the pumps.The operators generally understoodthat they should close one steamline to the feedwater heater aheadof the pump suction to control thewater temperature, but the reliabili-ty of this approach was still a con-cern.

To resolve this difficulty and toensure that the pumps never expe-rienced a flashing condition, an

10000

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14500

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0240 280 320 360

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NPSH IN FEET OF WATER

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■ ■■■ ■ ■ ■

HEAD vs NPSH945 GPM

2120 GPM

2635 GPM

3150 GPM

PUMP SIZE AND TYPE

ASSEMBLY NO. DATA BY

DRAWN BYFACTORY NO.

RPM

DATE

VOL. HYDR. LAYOUT

IMP. NUMBER VANES EYE DIA. EYE AREA MAX DIA. UNDERCOVER}FILEDTEST DIA. STGS.

R- 3262R- 3563R- 3134

35

6.06.06.0

33.015.615.6

642082642082642082

1ST333

8 x 10 x 14 A - 7STG HDB

803-E-1801

MA

MA

5775

21 MAY 81

CLEANING MOTOR H.P.INT. DRIVE

PEAK EFF.PEAK EFF.

T- 39005-3

BYRON JACKSON

Figure 2. Results of the NPSH test performed on Chalk Point’s boiler feed pumps.


“NPSH meter” was installed onthese units. The NPSH “meter” wasbasically a computer algorithm,installed in a plant’s computer, thatmonitored the suction conditions,that is pressure, temperature andflow. Based on temperature, thecomputer calculated the vapor pres-sure of the water. The net NPSHavailable was calculated by sub-tracting the vapor pressure fromthe suction pressure. The requiredNPSH was loaded into the comput-er from the pump manufacturer’scurve (Figure 2).

When the unit was in service, thecomputer continuously monitoredthe flow to determine the requiredNPSH. Thus, the NPSHA was con-tinuously calculated based on tem-perature and pressure. The comput-er subtracted the required NPSHfrom the available, and if there wasinsufficient margin, an alarm sig-naled the operator. The operatorcould then take whatever actionwas necessary to reduce the temper-ature or increase the suction pres-sure to solve the problem. This sys-

tem was tested after the pumpswere installed, and it performedperfectly.

Long Term ResultsThe four new pumps were

installed in 1981 and 1982. Sincethat time they have operated anaverage of more than 16 years.During this period they have hadroutine overhauls at intervals rang-ing from two to four years. Twospare volutes and rotating elementsare stocked at the plant. These aregenerally rotated into the pump tobe overhauled, and the old voluteassembly is rebuilt.

The total forced outage time dur-ing those 16 years with four pumpshas been 796 hours. The annualaverage forced outage rate is only12.5 hours per pump per year. Thehistorical average for the old pumpswas 389 hours per pump per year.That is a reduction of an impressive97% in the failure rate. This hasenabled the units to generate anadditional four million megawatt-hours of electricity during this 16-

year period. These pumps haveoperated well since they wereinstalled and continue to performwell.

SummaryAs the utility industry moves

toward deregulation, the impor-tance of having plants on-line tomeet electric demand will becomecritical. Reliable equipment isessential to this goal. This projectdemonstrates that with properattention to details and to forecast-ing equipment difficulties, a compa-ny can justifiably make a capitalinvestment that will optimize futuregeneration capacity. The role of thespecifying engineer should be toprovide the pump designer with thebest information possible on operat-ing conditions and allow the design-er to meet them. This case is aprime example of that philosophy. ■

Merwin W. Jones, PE, is the LeadEngineer for the Mechanical and CivilSection of the Potomac Electric PowerCompany.

? HANDBOOK PUMP AND SYSTEM TROUBLESHOOTING

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