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8.34 Pump Controls

F. B. HOROWITZ (1970) B. G. LIPTÁK (1985, 1995, 2005) S. BAIN (2005)

INTRODUCTION

This section describes the basic operation and controls ofpumps and pumping stations, while the next section (8.35)concentrates on the optimization of the unit operation ofpumping.

This section is divided into three main parts. The firstprovides a brief discussion of the processes into which thefluid is being transported by the pumps. The second partdescribes the types of pump designs, including centrifugal,rotary, and positive displacement designs and their basicmethods of controls. The third part discusses some aspectsof pumping system commissioning and operation.

The general discussion in this section is somewhat abbre-viated, because related topics are also covered elsewhere inthe handbook. Pumps, pumping stations, and metering pumpsare also discussed in Section 7.4, and variable-speed drivesin Section 7.10 in Chapter 7 of this volume. In addition,metering pumps are also covered in Section 2.14 in Chapter 2in the first volume of this handbook.

Some of the pumping system-related terms, abbrevia-tions, and conversion factors are described in Table 8.34a.

THE PROCESS

A pump is a liquid transportation device that must developenough pressure to overcome the hydrostatic and frictionalresistance of the process as it delivers the required fluid.These resistance components are unique characteristics of theprocess served and can be described by system curves. Thesystem curve of a process relates the pressure (head) requiredand the amount of fluid flow that is being delivered.

System Curves

The characteristics of the system that is served by a pumpor pumping station can be represented by a head-capacity

CentrifugalRotary

Reciprocating

TABLE 8.34a Pump Terms, Abbreviations, and Conversion Factors

*

Term Abbreviation Multiply By To Obtain

Length

L

ft 0.3048 m

Area

A

ft

2

0.0929 m

2

Velocity

v

ft/s 0.3048 m/s

Volume

V

ft

3

0.0283 m

3

Flow rate

Q

v

gpm 0.2272 m

3

/h

gpm 0.0631 l/s

Pressure

P

psi 6890 Pa

psi 6.89 kPa

psi 0.069 bar

Head (total)

H

ft 0.3048 m

NPSH

H

ft 0.3048 m

Output power (pump)

P

o

waterhp (whp)

0.7457 kW

Shaft power

P

s

bhp 0.7457 kW

Input power (driver)

Pi

kW 1.0 kW

Efficiencies (%)

Pump

E

p

— — —

Equipment

E

e

— — —

Electric motor

E

m

— — —

Utilization

E

u

— — —

Variable-speeddrive

E

v

— — —

System efficiencyindex (decimal)

SEI — — —

Speed

N

or

ω

rpm 0.1047 rad/s

Density

ρ

lb/ft

3

16.04 kg/m

3

Temperature

°

F —

°

C

* From Reference 1.

Flow sheet symbols

© 2006 by Béla Lipták

8.34 Pump Controls

2085

system curve (Figure 8.34b). The head at any one flow capac-ity is the sum of the static and the friction heads. The statichead does not vary with flow rate, as it is only a function ofthe elevation or back-pressure against which the pump isoperating.

The friction losses are related to the square of flow andrepresent the resistance to the flow caused by pipe and equip-ment friction. A system curve tends to be flat when the pipingis oversized and steep when the pipe headers are undersized.The friction losses also increase with the age of the plant.Therefore, the system curve for old piping tends to be steeperthan for new piping.

A generalized equation describing the system curve of aprocess is given below:

P

=

H

+

F

f

(

Q

x

)

8.34(1)

where

P

is the head pressure required to pump liquid into the process

H

is the static or elevation head of the process

F

f

is the friction factor of the process

Q

is the flow rate of the incompressible fluid

x

is an exponent that varies between 1.7 and 2.0, usually 2.0 is used

Static and Friction Pressures

Constant static head (

H

) isthe difference in pressure between the pump’s intake and itsdischarge at zero flow, if the discharge piping is full. Usuallyit is measured between the pump’s intake and the dischargeside of its check valve, thereby correcting the measurementfor any elevation differences between the two.

The constant static head can be the difference in elevationbetween the piping system’s intake and discharge (correctedfor liquid density if necessary). If the pump is discharging

into a pressurized system such as a potable water distributionnetwork or boiler, it will be the difference between the pumpintake pressure and the system’s pressure.

In any case, the constant discharge head is the pressurethe pump works against when the piping system is full andpressurized, but there is no flow through the pumps’ dis-charge pipes. As its name implies, it is comparatively con-stant. However, it may change, if for example the pump takessuction from a well and the well level drops, or if pumpingto a water storage tower and the water level in the towerrises.

Therefore, the constant discharge head is constant only inrelation to variations in one process variable: flow. If otherprocess variables (pressure, temperature, density, level) change,it will be affected; it is not constant.

The friction pressure component in Equation 8.34(1) is thepressure that is lost due to friction between the liquid andthe pipe. It includes losses from turbulence in bends and in theconversion of velocity pressure to static pressure in pipe expan-sions. However, for practical purposes they are combined intoa single term.

The value of exponent

x

in Equation 8.24(1

)

is not crit-ical. However, the controls can be improved by an accurateknowledge of the system curve, so the recommendation is tomeasure the value of

x

during commissioning.

Types of System Curves

Figure 8.34c illustrates the systemcurves of three different types of processes. Curve 1 corre-sponds to the closed-loop circulation of a fluid in a horizontalplane. Here, there is no static head component at all, and theparabola that describes the system starts at zero.

Curve 2 is the system curve for a condenser water circu-lation network. Here, a limited amount of static head ispresent, because the pump must return the water to the topof the cooling tower. This curve also illustrates that the fric-tion losses tend to increase when material builds up on theinside of the pipe, because it is no longer new. Curve 3 givesan example of a process dominated by static head. This isthe case when feedwater is being pumped into a boiler drum.This curve is flat and is relatively insensitive to changes insystem flow.

As will be discussed later in more detail, when thesystem curve is flat, there is little advantage to variable- ormultiple-speed pumping, and the usual response to systemflow variations is the stopping and starting of parallelpumps. Inversely, if the system curve is steep, substantialenergy savings can be obtained from the use of booster,multiple-, or variable-speed pumps.

Open and Loop-Type Systems

Hydraulic systems can be open(noncirculating) or loop-type, as illustrated in Figure 8.34d.Water supply and distribution systems in cities and buildingsare typically open systems, whereas hot-and-chilled-waterheating and cooling systems of plants are typically loop-typesystems.

FIG. 8.34b

The system head curve is the sum of the static head and the frictionlosses that have to be overcome in order to pump liquid into theprocess.

Capacity

Hea

d

Friction

losses

Total

static

head

System head curve


2086

Control and Optimization of Unit Operations

Hydraulic systems must also be evaluated as to whethertheir flow is restricted or unrestricted. Restricted-flow sys-tems are those that include valves that regulate the flowthrough the system. Hot-and-chilled-water systems, forexample, are restricted-flow systems, because manual orautomatic valves control their flow. Unrestricted-flow sys-tems include sewage and stormwater lift stations as well asthe pumping of municipal water into elevated storage tanks.

In actual systems, a single-system head curve may notexist. As illustrated in Figure 8.34e, what often exists is asystem head band. This is because the distribution of activeloads shifts the system curve within a wide band, as thisfigure shows.

FIG. 8.34c

System curves vary as the static-head components change in variousprocesses. (Adapted from Reference 1.)

System flow

Sys

tem

hea

d

Curve #1

(Horizontal liquid

circulation)

Curve #3

(Boiler)

Curve #2

(Cooling tower water)

System flow

Sys

tem

hea

d

Static riseto top of

cooling tower

Newpiping

Oldpiping

Friction headof condenserand piping

System flow

Sys

tem

hea

d

Boilerpressure

System head curve

Pipe friction

FIG. 8.34d

Pumping systems can be of the open or the loop type.

FIG. 8.34e

The system head drops when active loads are near the pump and itrises when the most remote loads are active.

4

Control Valves

Heatingor

coolingcoils

Loop type system

Open system

Heatexchanger

Pump

Pump

Storage tank

Water distribution systemto residences

Sys

tem

hea

d, F

T

System flow, GPM

100

0

10

20

30

40

50

60

70

80

90

0 200 400 600 800 1000

Active loadsnear pump

Uniform loading

Active loadsremote from pump

100GPM

100GPM

100GPM

100GPM

100GPM

100GPM

100GPM

100GPM

100GPM

100GPM

PHeat

exchanger

10 ft of pipe friction

between each load


8.34 Pump Controls

2087

PUMP TYPES

There are many types of pumps; however, almost all flow ispumped by just two types: centrifugal pumps and reciprocat-ing positive displacement (PD) pumps. As a result, this sec-tion will focus on these two types, and particularly on cen-trifugal pumps, which probably pump more flow than allother types combined.

Displacement and Centrifugal Designs

The two main pump designs are the positive displacement andcentrifugal pumps. Vertical centrifugal units can pump waterfrom depths up to 2000 ft (600 m), and horizontal units cantransport process fluids from clear water to heavy sludge atrates up to 100,000 gpm (6.3 m

3

/s). The centrifugal designsare of either the radial-flow or the axial-flow type.

Liquid enters the radial-flow designs in the center of theimpeller and is thrown out by the centrifugal force into aspiral bowl. A number of impeller designs are illustrated inFigure 7.4b in Chapter 7. The axial-flow propeller pumps aredesigned to push rather than throw the fluid upward. Mixed-flow designs are a combination of the two.

In positive displacement pumps, a piston or plungerinside a cylinder is the driving element as it moves in recip-rocating motion (Figures 7.4n and 7.4o). The stroke lengthand, thus, the volume delivered per stroke is adjustable withina 10:1 range. Rangeability can be increased to 100:1 by theaddition of a variable-speed drive.

The plunger designs are capable of generating higherdischarge pressures than the diaphragm types, because of thestrength limitation of the diaphragm. The strain on the dia-phragm is reduced if it is not attached directly to the plungerbut is driven indirectly through the use of a hydraulic fluid.Because solids will still settle in the pump cavities, thesedesigns are all limited to relatively clean services. For slurryservice, the hose-type design is recommended. This designeliminates all the cavities, although the seating of the valvescan still be a problem.

Pump Design Variations

If one attempted to give a more-or-less complete list of pumpdesigns, the list would include the following:

Progressing cavity pumps

are suitable for pumping liq-uids with entrained solids that would jam or abrade normalpump components, and liquids with low-shear requirements.The process industries use progressing cavity pumps forchemicals that may crystallize, polymers that need low-shearpumping, and liquids with some entrained solids. Progressingcavity pumps depend on the process liquid for lubrication,and they are easily damaged by low suction pressure. Theybehave like positive displacement pumps but are considerablyless accurate than reciprocating pumps, and their accuracydeteriorates with use, to the 2–5% level.

Peristaltic pumps

handle similar liquids as do the pro-gressing cavity designs but are generally more economicalin small (ml/s) sizes. The process industries use peristalticpumps in real-time wet chemistry analyzers that depend onaccurately metering small flows.

Gear, lobe, and vane pumps

deliver a relatively constantflow at constant speed with large changes in discharge pres-sure and, therefore, approximate the characteristics of posi-tive displacement pumps. These pumps cover the viscosityrange from less than 1 centipoise up to 500,000 centipoises.The usual application of this type of pump is for viscousliquids and slurries that are beyond the capabilities of cen-trifugal pumps.

Rotary screw pumps

are ideally suited for sludge andslurry services, and can safely pump high solids includinglive fish. A rotary screw pump design is also in developmentas an artificial heart. Rotary screw pumps behave similarlyto centrifugal pumps.

In

lift pumps

, compressed gas, usually air, is blown intothe bottom of a submerged updraft tube. The gas bubblesreduce the average density and, therefore, the hydrostatichead inside the tube. As the head is lower on the inside ofthe tube, the manometer effect induces the surrounding fluidto enter the updraft tube. This is a maintenance-free meansof lifting large volumes of slurries over low elevations but iscomparatively inefficient if one considers the compressorpower required.

The Archimedes Screw

is the earliest true pump and isstill in use. An advantage of a screw is that it naturallychanges its capacity to accommodate changing liquid levelswithout the complexity of a control system.

Air pumps

use compressed air or steam to displace accu-mulated liquids from tanks (Figure 8.6llll).

Diffuser micropumps

are generally specialized for use insmall-scale work.

Pumps specialized for high-vacuum work, such as

diffu-sion and cryogenic pumps.

CENTRIFUGAL PUMPS

Some centrifugal pumps use centrifugal force to throw liquidradially outward while others, such as propellers, use ascrew-type action that results in axial flow. Between thesetwo extremes, there is a whole continuum of impellers thatchange their pumping action from highly centrifugal radialto axial flow. The line separating centrifugal and axial-flowpumps is vague, and the behavior of these pumps is usuallydescribed by the same laws.

Figure 7.4b in Section 7.4 describes a number of impellerdesigns, including both radial and axial-flow types. Commoncharacteristics of centrifugal pumps are high efficiency (over90% in case of large pumps); they have only one movingpart (the impeller with bearings); they deliver smooth, steadyflow; they have a rangeability of about 4:1; and they are


2088

Control and Optimization of Unit Operations

relatively insensitive to air-locking, but are susceptible tocavitation.

Vertical centrifugal pumps can lift water from depths ofup to 2000 ft (600 m) and horizontal pumps can transportprocess fluids from clear water to heavy sludge at rates up to100,000 gpm (6.6 m

2

/s). The centrifugal pump is the mostcommon type of process pump, but its application is limitedto liquids with viscosities under 3000 centistokes.

Pump Curves

Figure 8.34f illustrates the typical pump curves of a singleimpeller pump.

Efficiency

The typical range of pump efficiencies is from60 to 85%. Pump efficiency is the ratio of the useful outputpower of the pump to its input power. Using the symbolsdefined in Table 8.34a, it is calculated (in both SI and USunits) as follows:

2

8.34(2)

8.34(3)

where

E

p

=

pump efficiency, dimensionless

P

i

=

power input, kW (kN·m/s)SP

=

specific weight of water, lb/ft

3

(kN/m

3

)

Q

=

capacity, ft

3

/s (m

3

/s).

H

t

=

total dynamic head, ft (m)

bhp

=

brake horsepower550

=

conversion factor for horsepower to ft lb

f

/s

Formulas

Table 8.34g provides a summary of the morecommon formulas that can be used in connection with pump-ing calculations.

EP Pp

i

t

i

= =pump output SPQHSI units( )

Ept= =

×pump output

bhp

SPQH

bhpU.S. customary

550( uunits)

FIG. 8.34f

Typical characteristic curves of a single-impeller centrifugal pump.

70(21)

60(18)

50(15)

40(12)

30(9)

20(6)

15(4.5)

10(3)

15(11.1)

5(3.7)

10(7.5)

300(0.019)

400(0.025)

500(0.032)

600(0.038)

700(0.044)

Capacity, GPM (m3/s)

NP

SH

Brake horsepower

NPSHR

Total head

Efficiency

30

40

50

60

70

Effi

cien

cy

(%)

Bra

ke

ho

rsep

ow

er (

kW)

1750 RPM

8" (200 mm) impeller

6" (150 mm) suction

5" (125 mm) discharge

To

tal

hea

d, f

t (m

)

TABLE 8.34g

Common Formulas Used in Connection with Pumping Calculation

1

Formula for Conventional Units SI Units

Head

H

=

psi

×

2.31/SG* (ft)

H

=

kPa

=

9.8023/SG* (m)

Output power

P

o

=

Q

v

×

H

×

SG*/3960 (hp)

P

o

=

Q

v

×

H

×

SG*/367 (kW)

Shaft power

Input powerEquipmentefficiency, %

P

i

–

P

s

×

74.6/

E

m

(kW)(Constant speed pumps)(Variable speed pumps)

P

i

=

P

s

×

100/

Em (kW)Ee = Ep × Em × 10−2

Ee = Ep × Em × Ev × 10−4

QD = design flow

Utilization efficiency, %QA = actual flowHD = design headHA = actual head

System Efficiency Index [see Eq. 8.34 (4)] SEI = Ee × Eu × 10−6

*SG = specific gravity

PQ H

Esv

p

=× ×

×SG

hp*

.( )

39 6P

Q H

Ev

ps

SGkW=

× ××

*

.( )

3 67

EQ H

Q HuD D

A A

=××

× 100


8.34 Pump Controls 2089

Characteristic Pump Curves The head-capacity curve isthe operating line for the pump at constant speed and impel-ler diameter. The characteristic curve of a pump describesthe variation of its discharge pressure with volumetric flow.The discharge pressure is the total of the velocity and staticpressures.

Three types of head-capacity curves are shown inFigure 8.34h, illustrating various relationships of capacity todischarge pressure. The capacity varies widely with changesin discharge pressure for all curves, but the shape of the curvedetermines the type of control that may be applied.

The shape of the head-capacity curve is an importantconsideration in pump selection. Curve 1 is referred as adrooping curve, curve 2 is called a flat curve, and curve 3would be considered normal. For on/off switching control,curves 1 and 2 are satisfactory as long as the flow is above100 gpm (6.3 l/s). Below this flow rate, curve 1 allows fortwo flows to correspond to the same head, and curve 2 maydrop to zero flow to obtain a small head increase. Both are,therefore, unstable in this region. Curve 3 is stable for allflows and is best suited for throttling service in cases in whicha wide range of flows is desired.

Matching the Pump(s) to the Process Figure 8.34i showsboth the head-capacity curve of a centrifugal pump and thesystem curve of a process. When such a system is uncon-trolled, the operating point of the system will be the point atwhich the pump and system curves cross each other.

If the process flow is controlled, a new system curve hasto be artificially generated. This can be done 1) by generatinga new system curve through the introduction of extra pressuredrop in a control valve, or 2) by changing the pump curvethrough changing the pump speed. Figure 8.34j illustratesthe case where a particular process flow is established bythrottling a valve and, thereby, changing the unthrottled(solid) system curve into a throttled system curve (dotted).

Throttling with a control valve makes the apparent pumpcurve steeper, so that it will cross the system curve at thedesired flow (operating point). This modification occurs atthe cost of introducing an artificial pressure drop of (Hp − Hs),which burns up pumping energy and, therefore, reduces oper-ating efficiency. In addition to that energy waste, the pumpwill also operate in a less-efficient region, as shown inFigure 8.34k by the throttled operating point #1 (72%) andthe unthrottled operating point #2 (80%).

Adjusting the Pump Speed

The affinity laws describe the relationships among changesin speed, impeller diameter, and specific gravity. With a givenimpeller diameter and specific gravity, pump flow is linearlyproportional to pump speed, pump discharge head relates to

FIG. 8.34hThe centrifugal pump’s curve can be drooping (1), flat (2), ornormal (3).

80(24)

60(18)

40(12)

20(6)

0 20(1.3)

40(2.5)

60(3.8)

80(5.0)

100(6.3)

120(7.6)

Flow, GPM (l/s)

Flat

(curve no. 2)

Drooping

(curve no. 1)

Pu

mp

dis

char

ge

pre

ssu

re

(in

fee

t o

r m

of

hea

d)

Normal

(curve no. 3)

FIG. 8.34i The system curve crosses the pump curve at the operating point.

FIG. 8.34j The addition of a control valve allows the control of flow at the costof added pressure drop (wasted energy).

Total static head

Friction and

minor losses

System head-capacity

curve

Pump head-capacity

curve

Hea

d

Capacity

Hea

d o

r p

ress

ure

Flow

2

1

3

Pump curve

Typical

throttled

operating

point

Thro

ttled

syst

em

Unthrottled

system

Natural

Operating point

Hs

Hp


2090 Control and Optimization of Unit Operations

(approximately) the square of pump speed, and pump powerconsumption is proportional to the cube of pump speed.

As shown in Figure 8.34l, one can plot the system curvesand the variable speed pump curves on a three-dimensionalplot (a-pressure, b-flow, c-speed). On such a plot the systemcurves form one surface (surface A), and the pump curves

form another characteristic surface (surface B). This thenillustrates how the intersection of surfaces A and B is theoperating line on which the variable-speed pump operates.(Section 8.35 will discuss the mathematical definition of thecharacteristic operating surfaces of pumps.)

Flow control via pump speed adjustment is less commonthan the use of throttling with valves, because most AC elec-tric motors are constant-speed devices. If a turbine drive isconsidered, speed control is even more convenient. However,the advent of the pulse-width modulated (PWM) adjustable-speed drive with sensorless flux-vector control has broughtadjustable-speed pumping into the mainstream of everydayapplications.

In order to vary pump speeds with electric motors, oneof the variable-speed drives described in Section 7.10 inChapter 7 should be used. The efficiency curves (wire-to-shaft efficiencies) of the various variable-speed drives thatare used on centrifugal pumps are shown in Figure 7.10z.

Variation of the pump speed generates a family of head-capacity curves, as shown in Figure 8.34m. If the impellerdiameter is constant, the volumetric flow through the pumpis proportional to its speed, and at reduced speeds, family ofspeed curves determines the flow rate (points 1, 2, or 3).

FIG. 8.34kThe throttled system not only wastes pumping energy through valvepressure drop but also operates at a less efficient point on the pumpcurve.

Hea

d o

r p

ress

ure

3

1

2

Thro

ttled

Unthrottled

Flow

60 65 70 75

75

78

78

80

80

100% speed

FIG. 8.34l The variable-speed pump operates on the line where the surface formed by the system curves (A) intersects with the surface formed by thepump curves (B).

10

8

6

4

2

0

0 0.2 0.4 0.6 0.8 1

Pre

ssu

reP

ress

ure

Flo

w

10

8

6

4

2

0

10

8

6

4

24

10

8

2

0

00Flow

Speed vs. pressure

Speed

0 0.2 0.4 0.6 0.8 100Speed

Flow vs. pressure

0 0.2 0.4 0.6 0.8 100

pressure

pre

ssu

re

Speed

Speed vs. Flow

Flow

The system curve

Flow (b)0 0.8

0.60.4

0.20

0

0.4

0.5

0.6

0.7

0.8

Speed

(c)Pressure

(a)

System

surface (A)

Pump

surface (B)

654321

1 0.8 0.6 0.4 0.40.60.8

1

0.20

0 0.20

Flow



Because the area of peak pump efficiency falls on aparabolic path, speed throttling will usually not reduce thepump efficiency as much as valve throttling (Figure 8.34k).This increases the total energy savings obtained from pumpspeed control. As shown in Figure 8.34n, when the flow isreduced from F1 to F2, instead of wasting the excess pumphead of (P1 − P2) in pressure drop through a valve, that pumphead is not introduced in the first place. Thereby, speed throt-tling saves the energy that valve throttling would have wasted.

When to Use Variable-Speed Pumps The shape of the sys-tem curve determines the saving potentials of using variable-speed pumps. As has been discussed, all system head curvesare parabolas (H ∼ Q2), but they differ in the steepness ofthese curves and in the ratio of static head to friction drop.As shown in Figure 8.34o, the value of variable-speed pump-ing increases as the system head curve becomes steeper.

Studies indicate that in mostly friction systems (such aszone 4 in Figure 8.34o), the savings represented by variable-speed pumping will increase with reduced pump loading.7 If,on the yearly average, the pumping system operates at notmore than 80% of design capacity, the installation of variable-

speed pumps can result in a payback period of approximately3 years.

The zones in Figure 8.34o are defined by the Ht /Hs ratio.The higher this number, the higher the zone number and the

illustrates how the Ht /Hs ratio is calculated. The shaded areasidentify the energy-saving potentials of variable-speedpumps. The values of Ht and Hs are identified on the basisof the average yearly flow rate (Fa) and determining its inter-sections with the pump and system curves. The larger the

t s

be and, therefore, the shorter the payback for the use ofvariable-speed pumps is likely to be.

FIG. 8.34mVariable-speed pump operation can be described by a family ofhead-capacity curves.

Pu

mp

dis

char

ge

pre

ssu

re

Flow

System

curve

(1)

(2)(3)

1,400 rpm

1,575 rpm

1,750 rpm

Pump

curves

FIG. 8.34oPumps and drives should be selected as a function of the steepnessof the system curve.

Zone 1

Zone

2

Zone

3

Zone 4

Pump curvesSystem head curves

Pump “B”

Pump “A”

Static

head

Head

(H)

Zone

1

2

3

4

Ht/Hs< 1.2

< 1.5

< 2.0

> 2.0

Flow (Q)

Pump sizes

Same

Various

Same

Various

Pump drives

Constant

Constant

One variable

All variable

FIG. 8.34nInstead of wasting the unnecessarily introduced pump energy, speed is reduced so that such energy is not introduced in the first place.6

F2

P2

P1

F1

Thro

ttle

d

Unthrottled

Flow

Hea

d o

r p

ress

ure

100% speed

67% speed

Speed, % Flow, %Horsepowerrequired, %

100

90

80

70

60

50

40

30

100

90

80

70

60

50

40

30

100

73

51

34

22

13

6

3


more justifiable is the use of variable-speed pumps. Figure 8.34p

shaded area in Figure 8.34p, the higher the H /H ratio will


Variable-Speed System Efficiencies The overall systemefficiency index (SEI) of a variable-speed pump installationis determined as follows:

8.34(4)

whereEp = the pump efficiency (%)Em = the motor efficiency (%)Ev = the variable-speed drive efficiency (%)Eu = the efficiency of utilization (%)

Figure 7.10z gave some typical wire-to-shaft efficiencyvalues for variable-speed drives (Ev). The efficiencies of thevariable-speed drives are represented as ranges, because eachdesign has a different efficiency. Figure 8.34q provides someindividual efficiency data for a number of variable-speeddrive designs.

At 50% of rated speed, the variable-speed drive efficiencycan be as low as 40% or as high as 70%, depending on the

design selected. Naturally, the less efficient the variable-speed drive, the less expensive it is. Table 8.34r providessome cost information for variable-frequency inductionmotor drives.

The efficiency of utilization (Eu) is an indicator of thequality of the overall piping and equipment design. For exam-ple, in the piping distribution system illustrated in Figure 8.34s,Qr might represent the required water flow and Hr therequired pressure head to transport Qr. Because three-wayvalves are used in this illustration, the actual flow (Qa) ismuch higher than the required flow, and the head required totransport this actual flow (Ha) is also greater than what isrequired.

Therefore, the efficiency of utilization, which is definedin Equation 8.34(5), will also be rather low for the designdescribed in Figure 8.34s.

8.34(5)

The numerical value of Eu can be accurately obtainedonly by testing. Once all four efficiencies are determined,they can be represented by a single system efficiency indexcurve (Figure 8.34t). By combining different pumps, motors,drives, and system designs, it is possible to arrive at a numberof SEI curves. The relative advantages of different devicesand designs can be evaluated quantitatively by comparingthese curves.

Types of Variable-Speed Drives Section 7.10 in this vol-ume provides a detailed description of the various variable-speed drive designs, their features including efficiencies, andcosts. Table 8.34u provides a summary of some of the mainfeatures of these variable-speed drives.

Two-Speed Pumps Two-speed motors can be utilized onsimple pumping systems in which accurate control of pumppressure is not necessary. Such motors should not be used insystems with high static head and low friction processes. Onthe other hand, for most friction systems, they can offer areasonably efficient and inexpensive alternative to variable-speed pumping. As shown in Figure 8.34v, standard two-speed motors are available with speeds of 1750/1150 rpm(29/19 r/s), 1750/850 rpm (29/14 r/s), 1150/850 rpm(19/14 r/s), and 3500/1750 rpm (58/29 r/s).

Cavitation

Cavitation can be a severe problem; it can quickly destroy apump. The cause of cavitation is that the pumped liquidflashes to vapor at one point in the impeller (where pressureis below the vapor pressure), and as the spinning impellerthrows the liquid and vapor outward, the formed bubblescollapse as the pressure rises above the vapor pressure. Whenthe collapsing bubbles come to the wall, they collide withextreme force. This gives rise to the characteristic sound of

FIG. 8.34pThe higher the Ht/Hs ratio, the more justifiable is the use of variable-speed pumps.

Friction

head

System head

curve

Static head

System flow

System flow

Hs

Ht

Sys

tem

hea

d

Design

head

100%0

HtHs

= 2.3

Pump head-capacity

curve

Static head

Hs

Ht

Sys

tem

hea

d

Design

head

100%0

HtHs

= 1.3

Pump head-capacity

curve

Average

yearly

flow rate

(Fa)

Friction head

SEI = × × × −( )E E E Ep m v u 10 6

EQ H

Q Hur r

a a

= 100



cavitation, and also to the consequent erosion, usually of theimpeller.

For this reason, it is important to operate pumps only intheir operating region where cavitation does not occur. Cav-itation usually occurs when the pump is delivering high flow,but some pumps cavitate at low flow too, when liquid recir-culating in the impeller passes through points of both lowand high pressures (Figure 8.34w).

When one does not have reliable data on the locations ofcavitation regions, it might be necessary to (very briefly)force the operating pump into cavitation intentionally. A prac-tical way to force a pump to cavitate is to run it at or abovedesign speed and then to restrict its intake gradually. Cavi-

tation has a very characteristic and memorable sound that hasbeen described as rocks rolling around inside the pump, orhammering on the impeller.

If a pump cavitates, one might try to eliminate or mini-mize the cavitation by improving inlet conditions. Just asrestricting the inlet will force a pump into cavitation, soremoving restrictions will often stop cavitation. Reducing theflow range of the pump usually helps also. If the pumpcavitates at high flows, try to avoid those flows, possibly bystarting a second pump sooner. If cavitation occurs at lowflows, one might turn off the pump at low flows, based onsome on/off flow control strategy.

Some impellers can accept a part, called an inducer, thatreduces the pump’s susceptibility to cavitation. An extremeoption is to inject a compressible gas into the impeller. Thisreduces pump efficiency and capacity, but it can eliminatecavitation, because the gas acts as a spring inside the liquid,absorbing the drastic localized pressure changes, and soavoiding the pressure extremes that the pure liquid wouldexperience.

However, cavitation is mainly a design concern andshould be dealt with during design by ensuring that amplepressure is available around the impeller.

Net Positive Suction Head

When liquids are being pumped, it is important to keep thepressure in the suction line above the vapor pressure of the

FIG. 8.34qVariable-speed drives in the 100-HP and larger sizes offer a wide range of efficiencies.9,10

90

80

90 10080

70

70

60

60

50

50

40

30

Dri

ve e

ffici

enc

y, %

(E v

)

% Rated speed

Pulse width

modulated

adjustable

frequency

Wound rotor

Eddy current

Hydraulic clutch

Variable voltage

Voltage source

adjustable

frequency

Regenerative

Static D. C.

90

80

90 10080

70

70

60

60

50

50

40

30

20

Effi

cien

cy

of

mo

tor

and

co

ntr

ol,

%

% Rated speed

1

2

3

4

5

6

Type of variable speed system

1 Energy recovery

2 Direct current

3 Current source variable frequency

4 Pulse width modulated variable

frequency

5 Eddy current coupling resistor

reactor (secondary)

6 Fluid coupling

TABLE 8.34r2004 Cost of Variable-Frequency Induction Motor Drives

Power Rating PWM Including Reactor Current-Fed ASCI

10 hp (7.5 kW)20 hp (15 kW)50 hp (37 kW)100 hp (75 kW)200 hp (150 kW)500 hp (375 kW)1000 hp (750 kW)

$1,700 $2,500 $5,300 $7,500$11,000

$8,000$12,000$18,000$32,000

Note: The reactor price adds 50–90% to the base price of thevariable-frequency drive electronics.



fluid. The available head measured at the pump suction iscalled the net positive suction head (NPSH).

A pump at sea level that is pumping 60°F water, whichhas a vapor pressure of hvp = 0.6 ft, and that is operatingunder a barometric pressure of 33.9 ft, has an available NPSH

(NPSHA) of 33.9 − 0.6 = 33.3 ft. If the impeller centerlineis 3.4 ft below the surface elevation of the water beingpumped and if the friction losses in the intake piping is 5.3 ftW.C, the NPSHA is 33.9 − 0.6 + 3.4 − 5.3 = 31.4 ft. Asshown in Figure 8.34x, the NPSHA increases with barometric

FIG. 8.34s The efficiency of utilization of this design is low, because the water that is bypassing the coils is being circulated unnecessarily.11

C

Coil

Coil

CoilQr

Qr

Qr

Remote building

Coil

Coil

CoilQr

Qr

Qr

Remote building

Primary

pump

Qa and Ha

Chiller

Central plant

Three-way

valves on

cooling coils

FIG. 8.34tOverall system efficiency index (SEI) curves describe the totalpumping efficiency as a function of load.1

1.00

0.80

0.60

60 80 100

0.40

40

0.20

200

0

System flow, %

Sys

tem

effi

cien

cy

ind

ex (

SE

I)

Minimum

flowDesign

flow

TABLE 8.34uVariable-Speed Drive Comparison

Drive Type

Efficiency (at 70% Speed) Turndown

Sizes (hp)

ComponentRequiring

Replacement(Frequency

in Yrs.)

Wound rotor regenerative

High (85%) 2:1* 25–500+ Brush (3–4)

Direct current High (80%) Unlimited 1–500+ Brush (1–2)

Variablefrequency

High (78%) 3:1 20–500+ —

Wound rotor Med. (60%) 2:1* 25–500+ Brush (3–4)

Eddy-current clutch

Med. (58%) 5:1 20–500+ —

Fluid coupling Med. (57%) 3:1 20–500+ —

Variable voltage Low (52%) Limited 10–100+ —

Mechanical Low (50%) 6:1 1–100 Belt or chain(1–3)

*Unstable below 50%.



pressure and with static head, and it decreases as vapor pres-sure, friction, or entrance losses rise.

Figure 8.34x also illustrates the difference between avail-able and required NPSH (NPSHA and NPSHR). AvailableNPSHA is the characteristic of the process and representsthe difference between the existing absolute suction head andthe vapor pressure at the process temperature. The requiredNPSHR, on the other hand, is a function of the pump design(Figure 8.34f). It represents the minimum margin betweensuction head and vapor pressure at a particular capacity that

is required for pump operation. If this minimum NPSHR isnot available, the pump will fail to generate the requiredsuction lift and the flow will stop.

The NPSHR curve in Figure 8.34f describes the requiredamount of static head at the pump inlet required to avoid adischarge pressure from dropping more than 3% over thezero-cavitation condition (source of definition: ANSI/HI9.6.1-1998, Paragraph 9.6.1.1). NPSHR is defined in absolutepressure, and typically for water at 0°C unless otherwisestated.

The available NPSH (the NPSHA) increases as 1) thebarometric pressure increases, 2) the static pressure of theliquid at the entrance of the impeller, and 3) all other suction-side pressures increase. The NPSHA decreases as 1) thevapor pressure of the liquid increases, 2) friction or entrancelosses rise, and 3) all other suction-side pressures decrease.

NPSH and Cavitation Traditionally, the NPSHR curve hasbeen taken to define the onset of cavitation, and designershave concentrated on ensuring that the NPSHR is met underall operating conditions (Figure 8.34x). However, the accu-racy of the NPSHR curve in defining the point when cavita-tion becomes significant is being questioned.

Indications are that cavitation actually begins sooner thanpreviously believed. Allan Burdis (chairman of the HydraulicInstitute’s NPSH Margin Committee) writes: “If you want tooperate cavitation-free, you need NPSH margin ratios (NPSHA/NPSHR) of 4 to 5.” Some evidence suggests that cavitationdamage at NPSHA = NPSHR is actually less than the damagewhen the NPSHA is higher. The Hydraulic Institute’s stan-dard report reads: “There are studies that show the maximumcavitation damage can actually occur at NPSHA values thatare twice the NPSHR or more for very high suction energypumps.”

Therefore, we might conclude that the science of cavita-tion prediction on the basis of NPSHA is still evolving, andif one wants to be positive, actual cavitation testing can benecessary.

Water Hammer

If a valve opening is suddenly reduced in a moving watercolumn, this causes a pressure wave to travel in the oppositedirection to the flow. When this pressure wave reaches a solidsurface (elbow, tee, and so on), it is reflected and travels backto the valve. If, in the meantime, the valve has closed, a seriesof shocks, sounding like hammer blows, results. An examplecan illustrate this phenomenon.

Assume that 60°F (20°C) water is flowing at a velocityof 10 ft/s in a 3 in. Schedule 40 pipe, and a valve located200 ft downstream is suddenly closed. The pressure rise andthe minimum acceptable time for valve closure can be cal-culated. If the valve closes faster than this time limit, waterhammer will result. In rigid pipe, the pressure rise (∆P) isthe product of water density (ρ in units of slugs/ft3), thevelocity of sound (c in units of ft/s), and the change in water

FIG. 8.34vSeveral speed combinations are available in two-speed pumps.1

FIG. 8.34wThe dashed segments of the characteristic pump curve show thezones where the regions of cavitation might exist.

Pump head-capacity curves

1750 rpm

Region of best efficiency

Best efficiency area

1150 rpm

850 rpm

550 rpm

20

18

16

14

12

120

10

100

8

80

6

60

4

40

2

200

0

To

tal

hea

d

Capacity

70

60

50

40

30

20

10

00 4 8 12 16 20

Pre

ssu

re

Flow

Cavitation region Characteristic



velocity (∆V in units of ft/s).12 Therefore, the pressure risecan be calculated as follows:

8.34(6)

In order to prevent water hammer, the valve closure time(t) must exceed the ratio of two pipe lengths (2L) divided bythe speed of sound:

8.34(7)

Therefore, in this example, the valve closure should takemore than 0.0823 sec.

The possible methods of preventing water hammerinclude (1) designing the system with low velocities, (2)using valves with slow closure rates, and (3) providing slow-closing bypasses around fast-closing valves, such as checkvalves.15

When water hammer is already present and the cause ofit cannot be corrected, its symptoms can be treated (1) byadding air chambers, accumulators, or surge tanks; (2) byusing surge suppressors, such as positively controlled relief

valves; and (3) when water flows are split or combined, byusing vacuum breakers to admit air and thereby cushion theshock resulting from the sudden opening or closing of thesecond split stream.

Pump Stations

When either the flow or the pressure requirements of theprocess are such that a single pump cannot meet them, pumpstations consisting of two or more individual pumps have tobe used. Multiple pumps operate in parallel and are used ifthe process flow rangeability exceeds the throttling capabilityof a single pump. Booster pumps are installed in series andare used to increase the total discharge pressure of the station.

Multiple Pumps in Parallel Individual centrifugal pumpshave a rangeability of about 4:1, which can be obtained byeither speed control or by discharge throttling. Pump turn-down can be increased by 1) bypassing the unwanted flow,2) turning the pump on and off, and 3) using multiple pumps.

When two or more pumps operate in parallel, the com-bined head-capacity curve is obtained by adding up theirindividual capacities at each discharge head, as illustrated in

FIG. 8.34x The available net positive suction head (NPSHA) increases with barometric pressure and static head, and decreases as vapor pressure,friction or entrance losses rise. (Adapted from Reference 3.)

Hea

d, i

n f

eet

of

liq

uid

Capacity

hb − hvp

hf + hihs

NPSH require

d

by pump Available NPSH

A

Static head

(+ hs)

B

hb

NPSH = hb + (hs − hvp) − (hf + hi)

NPSHA = hb − hs − hvp − (hf + hi)

Allowable

suction

lift

(−hs)

Allowable NPSH

(depends on impeller)

Pipe friction and

entrance losses

(hf + hi)Barometric

pressure

on liquid

surface

(hb)

Vapor pressure(hvp)

Pump

Liquidsurface

∆ ∆P c V= − = − −= =

ρ ( . )( )( )

,

1 937 4860 10

94 138 lbf/ft2 6653 8. PSI

t L c= = =2 2 200 4860 0 0823/ ( )( )/ . seconds



Figure 8.34y. The total capacity of the pump station is foundat the intersection of the combined head-capacity curve withthe system head curve. This point also gives the head at whicheach of the pumps is operating.

If the selection is to be very accurate, the head-capacitycurves should be modified by substituting the station losses(the friction losses at the suction and discharge of the indi-vidual pumps) so that the resulting “modified head” curvewill represent the pump plus its valving and fittings.

When constant-speed pumps are used in parallel, theadded increments of pumping can be started and stoppedautomatically on the basis of flow. As will be discussed later,a dead band is provided in these controls so that if a newpump is started at flow “x,” the flow will have to drop to, say,“x–5%” before that increment is stopped.

Booster Pumps When two or more pumps operate in seriesthe total head-capacity curve is obtained by summing up thepump heads at each capacity. When a booster pump is addedto a main fed by several parallel pumps, the total head-

capacity curve is obtained by adding the booster curve to themodified head of the parallel pumps at each capacity point.

Series pumping is most effective when the system headcurve is steep, such as in Figure 8.34z. With such mostlyfriction loads, series pumping can substantially reduce theoverpressure at low loads. Therefore, booster pumps or two-speed pumps can both be considered for the same kind ofsteep system curves. Multiple pumps in series are preferredfrom an operating cost point of view, but the capital costinvestment of a single two-speed pump is lower.

When constant-speed pumps are used, the booster pumpcan be started and stopped automatically on the basis ofpressure. In this case, an adjustable dead band is provided inthe pressure switch. As an example, the normal operatingpoint of the system can be point (1) in Figure 8.34z. As theload increases, the pump discharge head drops and when itreaches point (2) (the set point of the PSL), the booster pumpis automatically started. As soon as the booster is on, thesystem operates at a point to the right of point (3) until theload drops off again.

FIG. 8.34y Pump turndown and rangeability can be increased by operating twoor more pumps in parallel.

P4

P3

P2

P1

Check

valves

Modified head = P4 − P1Actual head = P3 − P2

a a

b bc c

Head-capacity curve,pump A

Head-capacitycurve,

pump B

Hea

dCapacity

Combinedhead-capacity

curve forpumps A

and Boperatingin parallel

(Q = QA + QB)

Maximum pointsof operation

(4)

Two-pumpoperation

One-pumpoperation

100%system

design head

Systemheadcurve

Onepump

Twopumps

(2)(5)

(3)(1)

Independenthead

System flowTwo pumps, each with a capacity of 50% design flow at 100% design head

Sys

tem

an

d p

um

p h

ead

Pump head-capacitycurves

FSHdead-band

50%designflow

100%designflow

FIG. 8.34zMultiple pumps in series are effective when the system head curveis steep. The two pumps illustrated by the lower graph are eachcapable of generating 100% design flow at 50% design head.1,2

100%designhead

Two-pumpoperation

Systemhead curve

Maximumpoints ofoperation

Onepump

One-pumpoperation

Twopumps

Pumphead-capacity

curves

(2)(5)

(4)(3)

(1)

System flowTwo pumps, each with a capacity of 100% design flow at 50% design head

Sys

tem

an

d p

um

p h

ead

PSLdead-band

100%designflow

Bypasses for

single-pump operation

a

a

b

b

c

c

d

d

e

eHead-capacitycurve, pump B

Head-capacitycurve, pump A

Hea

d

Capacity

Combined head-capacity curve

for pumps A and Boperating in series

(H = HA + HB)Checkvalves

PSL

On @ 2, Off @ 4

Independenthead

50% designhead



The booster pump stays on as the load drops below point(3) until the PSL turns it off at point (4). At this point, thesystem is automatically returned to the single pump operationat point (5). The dead band in the PSL prevents the on/offcycling of the booster pump at any particular load. The widthof the dead band is a compromise: As the band is narrowed,the probability of cycling increases, while the widening ofthe band results in extending the periods during which thebooster is operated unnecessarily. If the pumps are identical,their running times can be equalized by alternating them, sothat the pump with the higher running time will be the one thatis stopped first.

POSITIVE DISPLACEMENT PUMPS

Reciprocating pumps, such as the piston and diaphragmtypes, deliver a fixed volume of fluid per stroke. The controlof these pumps is based on changing the stroke length, chang-ing the stroke speed, or varying the interval between strokes.In all cases, the discharge from these pumps is a pulsed flow,and for this reason they are not suited to control by throttlingvalves. In practice, the volume delivered per stroke is lessthan the full stroke displacement of the piston or diaphragm.

This hysteresis is a result of high discharge pressures orhigh viscosity of the fluid pumped. Under these conditions,the check valves do not seat instantaneously. A calibrationchart must, therefore, be drawn for the pump under actualoperating conditions. A weight tank or level-calibrated tankis usually the reference standard. Because the discharge is apulsed flow (Figure 8.34aa), it must be totalized and dividedby the time interval to get average flow rate for a particularspeed and stroke setting.

Metering inaccuracy is approximately ±1% of the actualflow with manual adjustment and ±1.5% with automatic posi-tioning. Methods of stroke and speed adjustment are coveredin detail in Section 7.4 in Chapter 7, and other features ofmetering pumps are discussed in Section 2.14 in Chapter 2in the first volume of this handbook.

Reciprocating Pumps

Reciprocating pumps have a piston or plunger driving ele-ment inside a cylinder; the piston moves in and out, in areciprocating motion (see Figures 7.4n and 7.4o). Liquid issucked into the pump as the piston moves in, and is forcedout as the piston moves out. The liquid flows through one-way valves on the intake to allow flow only into the pump,and on the discharge only to allow it out.

The reciprocating piston action results in each strokedisplacing a positive, fixed volume of liquid almost regardlessof pressure, hence the name. Figure 8.34bb shows a typicalPD pump’s flow-to-pressure characteristic. The main featureof this characteristic is that it is essentially constant-flow:The pressure varies greatly while the flow varies very little.

Many reciprocating PD pumps allow the stroke length tobe adjusted, to change the pumping rate at constant speed,over a 10:1 range. Manufacturers offer higher stroke ranges,and theoretically they could adjust the stroke down to zero,but accuracy tends to decrease at low strokes, because offixed losses such as valve leakage.

PD pumps normally develop higher pressures than cen-trifugal pumps, and their flow is generally unaffected bydischarge pressure, which makes them well-suited to meter-ing applications. They can handle higher viscosity liquidsthan centrifugal pumps, and they can also handle slurries, butthey usually do not handle large entrained solids, as manycentrifugal pumps can (Figure 8.34cc).

Liquid compressibility can become a consideration atextremely high pressures, so that volume flow (gpm, m3/s)reduces slightly more than mass flow (lb/min, kg/s). PD pumpsare usually considerably less efficient than centrifugal pumpsexcept in high-head applications, so that large PD pumps arefar less common than large centrifugal pumps.

Some PD pump assemblies have multiple pumps on asingle shaft. This design is normally used to pump reagents

FIG. 8.34aaFlow characteristics of simplex and multiple plunger pumps.

Simplex pumpDuplex pumpTriplex pump

Average flow

for triplex pump

Averageflowfor

Duplexpump

Degrees motor shaft rotation

0 90° 180° 270° 360°

10094

63

50

31

Per

cen

t o

f m

axim

um

flo

w

Average flow

for simplex pump

FIG. 8.34bbThe characteristic curve of a positive displacement (PD) pumpoperating at constant speed and stroke.

Flow, GPM (l/s)

Pumpcurve

120(828)

100(690)

80(552)

60(414)

40(276)

20(138)

20(1.3)

40(2.5)

60(3.8)

80(5.0)

100(6.3)

120(7.6)

140(8.8)

160(10)

Dis

char

ge

pre

ssu

re P

SIG

(k

Pa)



or ingredients that need to be in a specific ratio. Operatorsadjust the individual pumps so that their volumes per strokeare precisely in the ratio required, then the pumps are drivenby a common shaft so that they all run at the same speedand, thus, deliver flow in a specific ratio.

Because flow is essentially independent of pressure, PDpumps that operate against a closed discharge (closed isola-tion valve, blocked pipe, and so on) can develop very highdischarge pressures that damage the equipment. For this rea-son, they often are provided with a pressure relief valve, orin applications with highly viscous liquids and slurries, pos-sibly with a rupture disc to relieve the excessive pressures,back to the pump inlet or the intake source.

A PD pump’s high pressure applies equally to the suctionside of the pump. Vacuum pressure can damage a PD pump,although this is only common in progressing cavity pumps,where an internal vacuum pressure tends to delaminate theflexible stator seal from the rigid stator support. Vacuumpressure can also damage piping and flexible metal couplings.

NPSH and Cavitation It is important to keep the NPSHabove 10 psia (69 kPa) or preferably above atmospheric.NPSH can be calculated as follows:

8.34(8)

whereP = feed tank pressure (psia)

Pv = liquid vapor pressure at pump inlet temperature (psia)Ph = head of liquid above or below the pump center line

(psid)l = actual length of suction pipe (ft)v = liquid velocity (ft/s)

G = liquid specific gravity

N = number of pump strokes per minuteC = viscosity (centipoise)d = inside diameter of pipe (in)

Flow Rangeability Speed adjustment is also an effectivemethod of adjusting a PD pump’s output, again generallyover a 10:1 range. One can combine both stroke and speedcontrol to achieve rangeability to 100:1, and add on- to off-time (mark-to-space ratio) control to further extend thatrangeability.

The PD pump’s ability to operate over a wide rangeabilitymay mean that only one pump is needed, which simplifiesoperation considerably. In process control, a PD pump’srangeability is important for chemical metering pumps, par-ticularly those used for pH control, where the need for adjust-ing reagents to 1 part in 100, 1000, or more is not uncommon.Rangeability of a hundred to a thousand is seldom reliablyachieved even with a PD pump, while it is impossible withcentrifugal pumps having a rangeability of only 4:1 or 5:1.

Calibration A PD pump’s check valves do not seat instan-taneously, so accurate operation needs to be based on a cal-ibration chart for the pump under actual operating conditions.Usually the reference standard for calibration is a calibrationcolumn or weigh tank that measures volume or mass. Cali-brate the pump’s flow by dividing the volume (or mass ifusing mass flow) by the time interval to get the average flowrate for a particular speed and stroke setting. Figure 8.34dd

FIG. 8.34cc When PD pumps are transporting slurries, it is advisable to userupture discs to protect against the development of excessive dis-charge pressures.

Variable

speed

coupling

M Rotary

screw pump

Spray

dryerFeed

tank

PSE

TT

TRC

NPSH = − ± −

+

P P PlvGN lvC

Gdv h 525 980

2

2

22

FIG. 8.34ddRatio and calibration controls for reciprocating pump. When levelhas reached LHS, the three-way valve returns to the “normal” pathand nitrogen enters the tank to initiate discharge. When level dropsto LSLL, discharge is terminated by venting off the nitrogen. CounterQQI is running while rising level is between LSL and LSH. Totalcount, when compared with known calibration volume, gives totalerror. Hand switch HS initiates calibration cycle by diverting thethree-way valve to the “calibrate” path.

FT

FY FRC

FY

FR PSV

Ratiostation

Pulsationdampener

Calibrationzone

Vent

Nitrogen

S

LSH

LSL

LSLL

FQI QQIP/D

HS

Normalpath

Calibrate

Toprocess

(Underpressure)

×



shows an automatic calibration facility that can be used formore advanced applications, or to calibrate the pump by handperiodically.

If the calibration is done on the suction side of the pump,this usually eliminates errors due to fluid compressibility,because close to the pump intake, the pressure is approxi-mately constant.

Pulsating Flow As illustrated in Figure 8.34ee, the recip-rocating pump produces a strongly pulsating flow.

One method of dampening pulsation is to use multiplecylinders, which, similarly to the way a car engine smoothesits pulsating thrust, smoothes out with multiple cylinders.The other option is to use pulsation dampeners, which areillustrated in Figure 7.4u in Chapter 7, Section 7.4.

Dampeners are usually connected into the piping nearthe pump discharge, and this is effective. However, perfor-mance can be improved by minimizing pipe friction lossesbetween the discharge and the dampener, and introducingsome friction in the line leading to the process user by apartially throttled valve to provide an acceptably smooth flow.

Valves Reciprocating PD pumps have check valves on theirintake and discharge, and this adds several requirements thatshould be considered in the overall design.

It is necessary to ensure that the discharge pressure isalways higher than the suction pressure. This can be guaranteedby placing a back-pressure regulator on the pump discharge,or possibly through the piping arrangement, as shown inFigures 7.4r and 7.4s in Chapter 7, Section 7.4. This is neededbecause otherwise the valves may open and allow flowthrough uncontrolled.

It is also recommended to orient the pump vertically, ornear-vertically. The check valves are usually seated, at leastpartially by gravity, so their orientation is important. It is also

desirable to strain or filter the pumped liquid to removeparticles that may jam the valves, and ensure that the pipingis scrupulously cleaned during and after installation.

Air-Locking and Cavitation When the purpose of a positivedisplacement pump is to meter the flow rate, certain precau-tions are needed. These include the removal of all entrainedor dissolved gases, which otherwise can destroy meteringaccuracy. Figure 7.4t in Chapter 7, Section 7.4, shows howentrained gases can be returned to the supply tank.

PD pumps airlock comparatively easily: If a large bubbleof gas accumulates between the intake and discharge valves,it can simply expand and contract through the pump strokes,and thereby stop the pumping action completely. This is aparticular problem if pumping liquids that give off gas, suchas sodium hypochlorite, which gives off chlorine; hydrogenperoxide, which gives off oxygen; or biological sludges,which give off methane, hydrogen sulphide, and other gases.

Therefore, one should ensure that the piping inherentlyintercepts and safely removes gases before they can enter thepump. If the removed gas is poisonous (chlorine and hydro-gen sulfide), it must be piped to a safe location.

In the case of hydraulic diaphragm pumps, the gas bubbleon either side of the diaphragm has the same effect. If thepump acts as though it is airlocked, but there is definitely nogas between the valves, there may be some on the hydraulicsside of the diaphragm.

PD pumps are unaffected by cavitation, because no sud-den collapse of the bubbles formed by cavitation is allowed.The piston strokes under complete control, the cavity (thebubble of vapor that is formed during cavitation) only col-lapses at the rate allowed by the piston. As a result, thedestructive velocities that can be achieved by cavitation in acentrifugal pump do not occur in a PD pump.

Chemical Metering Pump Operation A major operationalconsideration that applies specifically to chemical meteringpumps is the distance between the pump and the point wherethe pumped fluid enters the process. This results in dead time.Assume, for example, that this distance is 100 m and that theliquid’s speed in the pipe is 0.33 m/s.

If the pump is transporting a constant dilution water flow,it will take 300 sec from the time when the dilution is changedto the time the new dilution reaches the process. Conse-quently, dead time limits the best achievable control, and itis always desirable to minimize it.

One method to minimize this dead time naturally is toreduce the distance between the point of control and the pointof use. This may require a second pipe to carry the dilutionliquid. If using this arrangement, consider running the chem-ical pipe inside the dilution pipe, to provide inherent second-ary containment.

Another solution is to use feedforward based on a mea-surement that is in advance of the actual process requirement,by at least as much time as is the dead time.

FIG. 8.34eeMultiple pistons tend to dampen pressure fluctuations.

0 90

Sim

ple

x

180

Full stroke

Half stroke

180° 180°Suctionstroke

Average

3.17 × Average270 360 90 180 270 360

0 90

Du

ple

x

180

1st. head

1st. head

Drive rotation in degrees

3rd. head 2nd. head

2nd. head 1st. head 3rd. head

1st. head 2nd. head2nd. head

Average

1.58 × Average270 360 90 180 270 360

0 90

Tri

ple

x

180Average

1.06 × Average

0.90 × Average

270 360 90 180 270 360



Pacing the dilution or reagent flow to the chemical flowusually requires a separate metering pump, or possibly a flowcontrol valve. If diluting or charging to a number of processusers in a specific ratio, one can use fixed valves, such asneedle valves, to ensure the correct distribution ratio to indi-vidual dilution points.

CONTROL OF PUMPS

Capacity control of pumps must recognize the incompressibil-ity of liquids. For this reason, changes in the volumetric flowrate throughout the system occur simultaneously, and densityis constant at constant temperature, regardless of pressure.

Pump capacity may be affected by (1) a control valve inthe discharge of a pump, (2) one-off switching, (3) variation inthe speed of the pump, or (4) stroke adjustment of PD pumps.Flow control by on/off switching provides only zero or full flow,whereas the other control methods provide adjustable flows inthe system. The applicability of these four methods of capacitycontrol is a function of the pump type, such as centrifugal,rotary, or reciprocating. The possible types of capacity controlsfor the various pumps are summarized in Table 8.34ff.

On/Off Control

On/off switching is the most common capacity control in use.It has many disadvantages such as flow surges that oftenhinder processing, high friction losses, and high electricitypeak demand charges. However, on/off control is simple andcan be economical, as its consequences do not require rede-sign to accommodate the limitations of on/off control.

Pumps that are controlled only by starting and stoppingare said to be constant-speed (CS) pumps.

When pumping suspended solids or slurries, when thepump is stopped, a specific disadvantage of on/off control isthat the solids may settle out of the liquid and may not goback into suspension when the pump starts again. This cancause plugging.

This can also happen if pumping entrained oil and grease:The grease will float and may stick to the top of the pipe. In

these cases, an option is to keep the pump running: deliverflow through a circulating loop and control capacity with apressure-controlled bypass back to the feed tank. For exam-ple, one can provide intermittent flow to feed a centrifuge byopening an on/off valve by a cycle timer. Such a loop isshown in Figure 8.7j. The pressure-controlled bypass shownin this figure allows the normal pump flow to be maintainedin the loop, while the centrifuge feed valve is closed.

CS (on/off) pump operation is usually straightforward,except that the pump motor may overheat if it is started andstopped too frequently. Motors are usually rated for a maxi-mum number of starts per hour (sph), because bringing themotor up to speed involves higher currents than keeping it atspeed; perhaps ten times higher. Motor heating is propor-tional to current squared (I2R), so the heating during start-upmay be 100 times higher than normal.

Submersible pumps are usually rated for up to 15 sph,whereas large dry-pit pumps may be limited to 2–4 sph.Control systems must be designed to accommodate starts perhour limitations.

On/Off Level Control Figure 8.34gg illustrates the use oflevel switches for on/off pump control. The interlocks keepthe tank level between the settings of LSH and LSL. In thisillustration, the two-probe conductivity level switch operatesa relay. When conductive liquid reaches the upper LSH probe,the relay closes contacts H and I. At this point the pumpstarts, and although the level then drops below the LSHprobe, the pump keeps running because the holding contact(H) maintains the circuit.

When the level drops below the LSL probe, both theload (I) and the holding (H) contacts open, stopping thepump. When the level rises again, no action occurs when theLSL is contacted, because the holding contact is still open.However, when the level reaches the LSH, electrical contactis established, and the relay closes to repeat the pumpingcycle.

The bottom portion of Figure 8.34gg shows a simplepump starter circuit that is controlled by the three-positionhand-off-automatic switch. When the controls are in auto-matic (contacts 3 and 4 connected), the status of the interlockcontact I determines whether the circuit is energized andwhether the pump is on. The purpose of the auxiliary motorcontact M is to energize the running light R while the pumpmotor is on. The parallel hot lead to contact 6 allows theoperator to check quickly to see if the light has burned out.

The amount of interlocking provided for pumping systemsis usually greater than that shown in Figure 8.34gg. In additionto the overload (OL) contacts shown in Figure 8.34gg, con-trols often include safety overrides. These usually detectexcessive pressure or vibration, low flow, leaks (submersiblepumps), or motor winding temperature, and these overridesusually energize remote alarms.

Most pump controls include a reset button that must bepressed after a safety shutdown condition is cleared andbefore the pump can be restarted. If the same pumps can be

TABLE 8.34ffPump Control Methods

Possible Types of Controls

Method of Control On/Off Throttling

On/off switch Centrifugal, rotary,or reciprocating

Throttling control valve Centrifugal or rotary

Speed control Centrifugal, rotary, or reciprocating

Stroke adjustment Reciprocating



controlled from several locations, interlocks should be pro-vided to resolve conflicting requests. One method is to pro-vide an interlock to determine the location that is in control.This can be a simple local/remote switch or a more compli-cated system.

When many locations are involved, conflicts betweenrequests are resolved by interlocks that establish prioritiesbetween control locations on the basis of selecting the lowest,highest, safest choice of pump operation. In such installationswith multiple control centers, feedback should be provided,so the operators not only know the actual status of all pumps,but also are aware of any conflicting requests coming fromthe other operators or computers.

Multiple Speeds, Multiple Pumps When two-speed pumpsare used, added interlocks are frequently provided. One inter-

lock might guarantee that even if the operator starts the pumpin high speed, it will operate for 0–30 sec in low beforeadvancing automatically to high. This makes the transitionfrom off to high speed more gradual. Another interlock mightguarantee that when the pump is switched from high to low,it will be off the high speed for 0–30 sec before the low speedis engaged, to give time for the pump to slow down.

When several pumps are supplied from a common elec-trical feeder, it may be necessary to ensure that the feeder isnot overloaded by the inrush current, when pumps are startedsimultaneously. If this feature is desired, a 0–30 sec timedelay is usually provided between pump starts.

When two or more pumps are used, an alternator shouldbe interposed between the level switches and the pumpmotors. The alternator places the pumps in service in analternating sequence. Thus, if two pumps are used, each willhave half as many starts per hour. However, while one pumpis out of service, the other must pump all the flow so it willhave the same sph as a single pump.

Alternation tends to equalize pump running hours andreduces starts per hour. In Figure 8.34hh, a cooling waterreturn sump is illustrated. In this system, each pump isdesigned to handle the normal flow by itself, but both pumpsoperate together if abnormally high flows are required.

On/Off Flow Control Figures 8.34y and 8.34ii show a two-pump arrangement that responds to varying flow demandsmeasured on the discharge side of the pumps. Pump I nor-mally operates at point (1) (at 80 gpm and 36 ft, or 5 l/s and10.8 m). When flow demand increases to 120 gpm (7.6 l/s),the head drops to 22 ft at point (2), and FSH starts pump II.The combined characteristic gives 120 gpm at 40 ft (7.6 l/sat 12 m) at point (3). In this control scheme, a wide rangeof flows is possible without serious loss of discharge pressure.

On/Off Pressure Control As was illustrated in Figure 8.34z,a pressure switch may be used to start a spare pump in orderto maintain pressure in a critical service when the operatingpump fails. In this case, a low pressure switch would actuatethe spare pump, which is piped in parallel with the first pump.

FIG. 8.34ggOn-off pump-down interlocks often utilize two-probe conductivityswitches.

B

To pump-down

motor starter

3 Load

contact

(I)

A. C.

supply

Transformer

Relay

Holdingcontact

(H) Electrodes

LSH

LSL

M

LSL

LSH

Conductive liquid

1,000 gallon(3.785 l)betweensettings

Intermittent

discharge

at 100 GPM

(6.3 l/s)

Continuous feedat 75 GPM

(4.7 l/s)

H N

7

83

1

4

2 10

9

I

H–0–AOL’s

On front of panel

M

MR

5

6Auxiliary

motorcontact

Interlock contactfor automatic control

by LSH and LSL

11

FIG. 8.34hhOn-off level control of dual pump station.

Alternator

LSHH

LSH

LSL

M M

Discharge capacity is

5,000 GPM (315 1/s)

per pump

Continuous feed

at 4,000 GPM

(252 l/s)SUMP



A second possibility is to boost pressure, as shown inFigure 8.34jj. In this case, if pump I is normally operatingat point (1), when the discharge pressure rises to 50 ft (15 m)at point (2), the flow is reduced from 53 to 20 gpm (3.3 to1.3 l/s). At this point, the pressure switch (PSH) will startpump II and close the bypass valve. The system will now

operate at point (3) on the combined characteristic curve,delivering 60 gpm at 50 ft (3.8 l/s at 15 m) pressure.

Using PLCs The previously described simple control strat-egies can be easily implemented in hardwired controls. Witha programmable logic controller (PLC), it is easy to add manyother features beyond what is practical in hardwire, such asto equalize the pump run times better, rather than just alter-nating when it is necessary to stop or start a pump. Forexample, it is simple to stop the running pump that has loggedthe most hours or start the pump that has the lower sph.

One can also program the PLC to flush the piping peri-odically by running both pumps together. If it has been along time since both pumps ran together, and there is storagecapacity available, one can initiate the running of both pumpsby delaying the starting of one pump until more demand hasaccumulated and then run both pumps together. This strategyadds a pump start but is unlikely to result in an sph violationbecause it only happens if both pumps have not run togetherfor some time.

If a pump has not run for several days because demandis low, and there is some liquid available to pump, it is rec-ommended to run the pump regardless of demand, even ifonly for a few seconds. This strategy helps to keep the bear-ings lubricated and avoids developing flats on ball- or roller-bearings. It also adds starts, but again this only happens ifdemand is low, when starts are unlikely to be a consideration.

Starts per Hour Traditional SPH methods include the use ofreduced-voltage or soft motor starters to increase the allowablenumber of SPH. Other design solutions include the providingof enough storage capacity in the system to accommodate theSPH limitation. One generally tries to avoid the technique ofmeasuring the time between starts, and actively limit thesestarts, because this is an intrusive approach that is likely tointerfere with process performance.

Reduced-voltage starters increase the allowable numberof SPH, so they are a successful way of accommodating thesph limitation. However, when this means has beenexhausted, the next step is to provide storage capacity. Thesph limitation can be met by calculating the pumps’ minimumcycle time (minimum start-to-start time), which is given by

MCT = 4V/Q 8.34(9)

whereMCT is minimum cycle time in minutesV is the available storage volume Q is the design flow of the pump

Example: A 10,000 gpm pump is rated for up to fourSPH. Find the working volume needed to ensure that therating is not exceeded. Four SPH corresponds to an MCT of15 min. Therefore,

= 37,500 gallons.

FIG. 8.34iiOn-off flow control can be used with parallel pumps.*1.0 ft of water = 2.98kPa.

FIG. 8.34jjOn-off pressure control of pumps.*loft of water = 2.98kPa.

M M

Pump

IPump

II

FSH

Set

at

120 GPM

(7.6 1/s)

50(15)40

(12)30(9)20(6)10(3)

0 20(1.3)

60(3.8)

100(6.3)

140(8.8)

180(11.3)

220(13.5)

260(16.4)

80

(5.0)

40

(2.5)

Flow, GPM (1/s)

Pu

mp

dis

char

ge

pre

ssu

re(i

n f

eet

or

m o

f h

ead

)*

Single

pump curve

Combined

pump curve

(5)

(1)

(4)

(3)

(2)

FSH dead-band

M

Pump I

M

Pump II

PV PSH

110(33)100(30)90

(27)80

(24)70

(21)60

(18)50

(15)40

(12)30(9)20(6)10(3)

0 70(4.4)

50(3.2)

30(1.9)

10(0.6)

60

(3.8)

40

(2.5)

20

(1.3)

Flow, GPM (l/s)

Pu

mp

dis

char

ge

pre

ssu

re(i

n f

eet

or

m o

f h

ead

)*

(2) (3)

(1)

Singlepumpcurve

Combined

pump curve

V = ⋅15 10 0004,



Modulating Control

The flow from a pump or pumping station can be controlledby throttling either the forward flow or the bypass flowaround the pump. Capacity control by valve throttling worksessentially by throwing away what is not needed. The unnec-essarily introduced pumping energy is wasted by eitherdiverting the flow that is not needed through a bypass, or byrestricting the pump discharge.

Pumps are usually powered by induction motors, whichfor a long time have been constant-speed devices. As a result,it was difficult or impractical to adjust pump speed, and alter-native ways of controlling the pump’s capacity were required.Modulating valves made CS pump capacity control practical,but they wasted energy. The development of reliable adjustable-speed (AS) induction motor drives has changed this situation,although modulating valves are still used.

Bypass Valves With a bypass valve, the pump deliversessentially a constant flow, and the bypass valve returns what-ever the process does not need back to the pump inlet. Theprocess determines by the opening of the flow control valve.For example, if the pump is supplying heating water to areactor, when more heat is called for, the bypass valve isthrottled to close, to increase heat to the reactor.

Bypass valves can work with both PD and centrifugalpumps, and they have other advantages, such as 1) Whenpumping slurries, greases, or mixtures that may separate,solidify, or coagulate, the bypass keeps the flow moving andso tends to keep the liquid homogeneous. 2) When pumpingheating water, the water in the piping tends to lose heat. Abypass keeps the flow constant, which tends to ensure aconsistent temperature to the process. 3) When pumping liq-uids with entrained gases, keeping the flow moving tends toavoid accumulation of larger gas bubbles that can airlock thesystem. 4) The heat gain in the pumped liquid stays constant.This is particularly important if pumping liquids near theirvapor pressure.

Throttling Valves Throttling restricts the pump’s dischargeflow. When pumping incompressible (liquid) flow, throttle thedischarge, but when gas (compressible) is transported, throttlethe inlet. With a liquid, the problem with inlet throttling is thatit causes cavitation, which cannot happen with a gas. With agas, the advantage of inlet throttling is that the throttled gasexpands into the blower at a reduced density (which reducesits power consumption). However, this cannot happen withan incompressible liquid — incompressible also impliesinexpandable. So, the arrangement of liquid discharge throttling,and gas inlet throttling, makes the best of both possibilities.

Essentially, throttling changes the friction factor of thesystem curve. It works well with centrifugal pumps by shiftingthe pump’s operating point, but does not with PD pumpsbecause their characteristic curve is very steep (Figure 8.34bb),and therefore the discharge pressure just increases to forcethe positive displacement flow through the valve.

In Figure 8.34kk, design point (1) is near the maximumefficiency of the pump. Therefore, when throttling to point(2), the efficiency will drop.

For good controllability, the control valve should be sizedto pass the design flow with a pressure drop equal to thesystem dynamic friction losses excluding the control valvebut not less than 10 psi (70 kPa) minimum. (For more detailson assigning sizing pressure drops to control valves, refer toSection 6.15 in Chapter 6.)

The control of flow by varying the pressure drop acrossthe valve is illustrated in Figure 8.34kk. Here, when the flowis throttled from point (1) at 73 gpm (4.5 l/s) to point (2) at15 gpm (0.95 l/s), the differential pressure across the controlvalve increases from 15 PSID (100 kPa) to 50 PSID (350 kPa).

However, one should be careful not to run the pump atflows low enough to overheat the liquid, or vaporize it,because vaporization can cause the pump to cavitate. Oneshould use a heat balance on the pump to calculate the min-imum flow needed through the pump to prevent vaporization.To be conservative, assume that all the motor’s power isconverted into heat. If this minimum flow was calculated tobe 20 gpm (1.3 l/s), then size the PCV in the bypass to pass20 gpm with a corresponding set pressure of 63 psi (435 kPa).

Figure 8.34ll shows a typical flow control loop with apressure-controlled kick-back bypass. The rangeability of thecontrol valve is assumed to be 25:1 (see Section 6.7). Thus,if the maximum flow required is 70 gpm (4.4 l/s) through theflow control valve, then the minimum controllable flowwould be about 3 gpm (0.2 l/s). If control of lower flows isrequired, then install a second, smaller flow control valve inparallel with the first, and use it to control these lower flows.

Throttling Valves Waste Energy If there is no throttlingvalve in the system, a constant-speed pump will always oper-ate at the intersection of its characteristic curve and the system

FIG. 8.34kkThrottling control of centrifugal pump.

70(483)

60(414)

50(345)

40(276)

30(207)

20(138)

10(69)P

um

p d

isch

arg

e p

ress

ure

PS

IG (

kP

a)

10(0.6)

80(5.0)

70(4.4)

60(3.8)

50(3.2)

40(2.5)

30(1.9)

20(1.3)

Flow, GPM (l/s)

Pumpcurve

30%

effi

cien

cy

50%

effi

cien

cy

(2)

(1)15 PSID

(103.5 kPa)

control valve

drop

15 PSID system

friction loss

13 PSID

(89.7 kPa)

Static loss

System friction plus

static loss curve

50 PSID(345 kPa)controlvalvedrop



curve. If the actual system curve has less slope than wasdesigned for, the pump will deliver more flow than wasintended. For example, in the case of the pump and systemcurves shown in Figure 8.34mm, the designers expected onesystem curve (solid line), but the actual system curve (dashedline) turned out to be flatter. Consequently, the actual oper-ating point will be at point (2) instead of point (1), andtherefore the actual flow will be higher, and the pressurelower, than the designers intended.

If a throttling valve on the pump discharge controls theflow, the pump will operate at point (1), and the valve will

burn up the differential pressure between points (1) and (3).This drop represents wasted power:

8.34(10)

wherePw is the wasted power (watts, W)p is the differential pressure across the valve (Pa)

Q is the flow through the valve (m3/s)

If the valve throttles the flow to 50%, as in points (2) to(4), the energy wasted in the form of valve pressure dropincreases, because it becomes the difference between points(4) and (5). This illustrates that throttling always wastesenergy. In addition, throttling introduces another source ofenergy wastage, because it almost always also reduces thepump’s efficiency (Figure 8.34k).

In Figure 8.34j, the useful pumping pressure is identifiedas Hs, and the actual pressure of the throttled system is givenas Hp. The (Hp − Hs) difference identifies the energy wastedthrough throttling. However, this is only part of the totalwaste, because moving the operating point from (2) to (1)also reduces the pump efficiency from 81 to 71%. As a result,not only is power lost to throttling, but power is also lost toreduced pump efficiency.

One can use Equation 8.34(10) to quantify the loss, bytaking the example in Figure 8.34kk. If working with metricunits, when the flow is throttled from point (1) at 4.5 l/s to point(2) at 0.95 l/s, the pressure across the valve increases from 100kPa to 350 kPa. At point (1), the flow is 4.5 l/s (4.5 × 10–3 m3/s)and required pressure is 190 kPa (4.5 × 103 Pa), giving arequired power of

However, the consumed power, including the valve’spressure drop, is

Therefore, the control’s efficiency at point (1) is = 65%We can now compare that result with point (2), which

has a flow of 0.95 l/s and required pressure of only 95 kPa,giving a required power of

The consumed power, including the valve’s pressuredrop, is

Therefore, the controls efficiency at point (2) is =20%Figure 8.34kk does not provide the full pump efficien-

cies, but point (1)’s efficiency is likely to be about 60%, and

FIG. 8.34llThrottling control with pressure kickback.

FIG. 8.34mmIllustration of the consequence when the assumed and actual systemcurves are not the same.

FT

FRC

PCV

M

Actual system

head curve

Point 5

Point 3

Point 2

Point 1

Point 4

50%

design

flow

100%

design

flow

System flow

Sys

tem

an

d p

um

p h

ead

Design systemhead curve

Designhead

Overpressurewith constantspeed pump

Pump head-capacity curve

Pw p Q= ⋅

Pw = × ⋅ × =−4 5 10 190 10 8503 3. w

Pw = × ⋅ × =−4 5 10 290 10 13003 3. w

8501300

Pw = × ⋅ × =−0 95 10 95 10 903 3. w

Pw = × ⋅ × =−0 95 10 440 10 4203 3. w

90440



point (2)’s about 15%. Including these values gives the over-all efficiencies of 40% (65 × 60%) for point (1) and 3% (20 ×15%) for point (2).

Capacity Control by Stroke Adjustment The flow gener-ated by a PD pump is the product of the volume per strokeand the number of strokes per second. Therefore, one canadjust the PD pump flow by either adjusting the volume perstroke or the strokes per second (the speed). The volume ofa piston is its area multiplied by the stroke length. Becausethe stroke length is readily adjustable, stroke control refersto the pump’s flow adjustment method of modulating itsstroke length.

The stroke can be adjusted by hand or automaticallywhile the PD pump is operating at constant speed. Handadjustments are currently more precise, usually rated at about1% error, compared to 1.5% error when automatic adjustmentis used. While manual setting is more accurate, automaticadjustment by closed-loop control will give better perfor-mance overall.

The range of flow control by stroke adjustment is0–100%. However, in order to maintain accuracy, the prac-tical range is 10–100% of maximum flow. The flow is relatedto stroke length through system calibration.

In chemical metering pump applications, it is often thecase that the automatic controls modulate the pumps’ speeds,while the operators adjust the strokes by hand. The reasonfor this technique is to use stroke adjustment to compensatefor factors that are hard to measure and remain constant overlong periods, such as the reagent concentration, that maychange only from delivery to delivery. In such configurations,the feedback controls automatically adjust the speed to com-pensate for factors that change often and can be measured,such as the process flows or the reagent deterioration overtime, possibly caused by temperature and so on.

Capacity Control by Speed Adjustment The cost of reli-able and efficient adjustable-frequency drives (AFDs) hasdropped rapidly in the past decade. As a result, adjustable-speed pumping is preferable today for both centrifugal andPD pumps. The main advantage of adjustable speed pumpingis that it is efficient. It is efficient in two ways: 1) Rather thanwasting energy (as do valves), speed control avoids introduc-ing unnecessary energy in the first place, and 2) A modulatingvalve almost always moves the pump’s operating point awayfrom its best efficiency point (BEP). Speed control alsomoves the pump away from its BEP, but not nearly as much.

The basic concept is that a pump’s speed controls itsdischarge flow: increase speed to deliver more flow, andreduce it to deliver less.

Multiple-Pump Controls

Distribution Controls From a control quality point of view,distribution controls have already been discussed in Section 2.23

in Chapter 2 and Section 8.28. Here, their description willbe from the perspective of the performance of the pumpstation. The overall efficiency of the water distribution sys-tem, which was shown in Figure 8.34s, is described inFigure 8.34nn by curve 1.

A substantial increase in efficiency (reduction in oper-ating cost) can be obtained by replacing the three-way valvesin Figure 8.34s with two-way ones and by replacing thesingle large pump with smaller ones. Figure 8.34oo showssuch a system. Here, a small and a large primary pump areprovided at the main supply point in the central plant, anda small and a large booster pump are furnished in each ofthe user buildings.

When the load is low, the small pumps are operating;when it is high, the large pumps take their place. The mini-mum flow requirements of the chiller are guaranteed by abypass valve, and the chilled water makeup into the recircu-lating loop of each building is under temperature control(TC). The resulting improvement in overall efficiency isshown by curve 2 in Figure 8.34nn.

FIG. 8.34nnOverall efficiency is maximum if two variable-speed pumps are usedand is minimum with a constant volume installation.4

System flow, % full load

Ove

rall

effi

cien

cy,

%

80

70

60

50

40

40

30

20

20

10

00

80 10060

Large

pumps

Small

pumps

Curve (3)

Variable speed pumps(Figure 8.34pp)

Curve (2)

Constant speedpumps

(Figure 8.34oo)

Curve (1)

Constant volumesystem

(Figure 8.34s)



The highest overall efficiency can be obtained throughthe use of variable-volume load-following. Figure 8.34ppillustrates such a system, utilizing variable-speed pumps intwo sizes. In this system, all waste is eliminated except what

is generated by the small minimum flow bypass around thechiller, which is guaranteed by a small constant-speed pump.The resulting increase in overall efficiency is illustrated bycurve 3 in Figure 8.34nn.

FIG. 8.34ooSupply-demand matching can be achieved using constant-speed pumps of different sizes.2

C

Coil

Coil

Coil

Remote building

Chiller

Central plant

Two-way

control valves

on cooling coils

CPrimary pumps,

one small and one large

constant speed pump

Chiller bypass

valve for

minimum flow

TC

C C

Coil

Coil

Coil

Remote building

CC

TC

Secondary booster pumps

in each building,

one small and one large

constant speed pump

Return temperature

control valves

FIG. 8.34ppVariable-volume water distribution systems provide maximum efficiency.4

Coil

Coil

Coil

Remote building Remote building

Chiller

Central plant

Two-way valves

on cooling coils

Primary pumps. One

large variable speed

and one small

variable speed pump

Chiller bypass

pump started by

FSL to guarantee

minimum flow

Secondary pumps in

each building.

One large and one small

variable speed pump

V V

Coil

Coil

Coil

V V

C

FSL

V V



Starting and Stopping Pumps When the running adjust-able speed pumps are near their maximum speeds and whenmore flow is needed, the capacity of the pumping stationmust be increased by starting another pump. Similarly, whenall pumps are at a low speed and less flow is needed, a pumpmust be stopped.

It is necessary to know the maximum pump speed a pumpcan run at and still deliver zero flow (called omega zero, ωo),the pressure up to which the flow remains zero (shut-offpressure). This is the pressure that a starting pump mustovercome before it can start delivering flow (superimposedback-pressure), and the flow delivered by an adjustable-speedcentrifugal pump, which is running against the back-pressure,which is superimposed by the other operating pumps.

The subject of determining the above values for particularpumps and processes and to use these values in developinga detailed strategy for starting and stopping individual pumpsis beyond the scope of this section, but if the reader needssuch information, it can be found in Reference 15.

When to Start or Stop a Pump Accumulated running hoursshould be recorded on an elapsed time meter (ETM), becausepumps are subject to wear while running, so recording theirrunning hours can help distribute wear among them evenly,and it provides a guide to planned maintenance.

The time since the last stopping of the pump, called idletime, also should be recorded. During idle time, pump bear-ings lose lubrication, and they can be flattened out of round,so long idle periods are undesirable. Recording the idle timehelps deal with this. A pump accumulates idle time whilestopped, in the same way that it only accumulates runninghours while running.

In critical applications consider always keeping an extrapump running at the same speed as the other pumps, so thatif a pump fails it is only necessary to speed up the runningpumps. One of the pumps should be stopped when the speedof the operating pumps has dropped to its low limit for over30 sec.

The operator or the control system should initiate a startimmediately when a running pump fails, unless a start isalready in progress. A pump should also be started when therunning AS pumps have been at their maximum speeds (ωm)for over 30 sec. Similarly, if an operating pump has traveledon its characteristic curve to the point where cavitation starts(Figure 8.34w), an additional pump should be started.

Start a pump if its idle time indicates that its bearingswould benefit from being rotated. Preferably do this whenthe pumps are running around the middle of the speed range,so that a normal start or stop is unlikely to occur at the sametime. Probably also stop another pump at the same time, toavoid a surge or having too much running capacity.

Selecting the Pump to Start or Stop When it is time tostart a pump, start the AS pump that has accumulated thelongest idle time. If all the AS pumps are running, start the

CS pump that has the longest idle time. This strategy avoidsaccumulating idle time, consistent with not adding any extrastarts.

When it is time to stop a pump, stop the running CSpump that has accumulated the most running hours. If all CSpumps are stopped, stop the AS pump that has accumulatedthe most running hours. This strategy tends to equalize thepumps’ running hours, but it is not extremely successfulbecause equalizing hours has the lowest priority— everythingelse takes precedence.

A highly successful strategy is to start the pump withfewest hours, stop the pump with the most hours. However,this can allow equipment to stand idle for long periods, par-ticularly when other equipment is brought back into serviceafter a long downtime.

Starting and Stopping When the pumps are of equal sizeand a pump is being started, accelerate it up to the omegazero speed (ω 0) quickly on its AFD’s acceleration ramp.Then, gradually increase the flow through the startingpump, and reduce it through the running adjustable-speedpumps, until they are all running at the same speed andpassing a similar flow. Similarly, when stopping, controlthe pump’s deceleration down to ω0, while accelerating upthe running pumps. When the stopping pump reaches ω0,stop it quickly.

When starting an AS pump, it is necessary to know howmany AS pumps are running. Next, determine the startingpump’s ω0. If necessary, perhaps because as-commissionedpump measurements may not be available when the controlsare configured, use the characteristic and system curves toestimate ω0 against n. Start the pump and accelerate it up toω0 quickly.

When the starting pump reaches ω0, start a timer, thetransition timer, to bring the starting pump into action. Letthe duration of the timer be T sec and the elapsed time sincethe start be t sec.

Depending on the system, the timer’s duration may befaster or slower; it is usually not critical, provided it is slowenough that the pumps can follow the flow changes itrequires, and fast enough to ensure it will have timed outbefore there is any need to start another pump.

While the transition timer is running (0 < t < T), calculatethe starting pump’s flow increase, and the running pumps’flow decrease, to pump the instantaneous flow required bythe control loop output, Q, throughout the start transition.Then convert these flows to pump speeds, as follows:

When the transition timer has timed out (t = T), thestarting pump has joined the running pumps.

Update the count of the number of running pumps, n. When stopping an AS pump, the principles are the same

as was for starting an AS pump, except in reverse: deceleratethe stopping pump to ω0 while accelerating the runningpumps, so that total flow is controlled while it also transitionssmoothly from the stopping pump to the running pumps.



CONCLUSIONS

This section covered the basics of pump control, while thenext section will concentrate on the optimization of this unitoperation. Pumping controls are a prime example of appli-cations where it is essential to fully understand the person-ality of both the process and the pumping equipment used,before a successful control system can be designed.

Another unique characteristic of the pumping process isthat over the life of the plant, the operating cost of a pumpis much greater (sometimes a hundred times greater) than thefirst cost of the pumping equipment. It is for this reason thatgood process controls and optimization can have much higherreturns when operating pumping stations than on other unitoperations.

The goal of a well-designed pumping control system isgood supply–demand matching, which will not only loweroperating costs, but also reduce maintenance and cycling.The full automation of pumping stations — including auto-matic start-up and shutdown — not only will reduce operatingcosts but will also increase operating safety as human errorsare eliminated.

References

1. American Society of Heating, Refrigerating, and Air-ConditioningEngineers (ASHRAE), “Centrifugal Pumps,” in ASHRAE Handbook,2003 Equipment Volume, Atlanta, GA: ASHRAE, 2003, Chapter 39.

2. Tchobanoglous, G., Wastewater Engineering, Metcalf and Eddy, Inc.,New York: McGraw-Hill, 1981.

3. Karassik, I., Centrifugal Pumps, F. W. Dodge Corp., 1960.4. Rishel, J. B., “Water System Head Analysis,” Plant Engineering,

October 13, 1977.5. Conzett, J. C., “Adjustable-Speed Drives,” Bulletin D-7100, Reliance

Electric, July 1981.6. Lipták, B. G., “Save Energy by Optimizing Your Boilers, Chillers,

and Pumps,” InTech, March 1981.7. Langfeldt, M. K., “Economic Consideration of Variable Speed

Drives,” ASME Paper 80–PET–81, 1980.8. Liu, T., “Controlling Pipeline Pumps for Energy Efficiency,” InTech,

June 1979.9. Merritt, R., “What’s Happening with Pumps,” Instruments and Con-

trol Systems, September 1980.10. Schroeder, E. G., “Choose Variable-Speed Drives for Pump and Fan

Efficiency,” InTech, September 1980.11. Rishel, J. B., “Wire to Water Efficiency of Pumping Systems,” Central

Chilled Water Conference, 1975, Purdue University.12. Baumeister, T. (Ed.), Mark’s Standard Handbook for Mechanical

Engineers, 8th edition, New York: McGraw-Hill, 1978, p. 3.69.13. Shinskey, F. G., Energy Conservation Control, New York: Academic

Press, 1978.14. Systecon (Division of CEC), Bulletin No. 10-320-1.15. Karassik, I., “What are the Characteristics of the Unknown Pump?”

Center for Professional Advancement.

16. Bain, S., “Variable-Speed Pumping Requires Careful Study of Con-ditions,” WEFTEC 65th Conference, Collection Systems Proceed-ings.

Bibliography

Avallone, E.A., and Baumeister, T., Mark’s Standard Handbook forMechanical Engineers, New York: McGraw-Hill, 1996

Baumann, H. D., “A Case for Butterfly Valves in Throttling Applications,”Instruments and Control Systems, May 1979.

Baumann, H. D., “Control Valves vs. Speed Controlled Pump,” Texas A&MSymposium, 1981.

Baumann, H. D., “How to Assign Pressure Drop Across Liquid ControlValves,” Proceedings at 29th Annual Symposium on Instrument Engi-neering for the Process Industry, Texas A&M University, January,1974.

Bose, B. K., Power Electronics and AC Drives, Englewood Cliffs, NJ:Prentice Hall, 1986.

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