Pumps Specification 80031E

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RADIAL, MIXED AND AXIAL-FLOW PUMPS.SIZE ESTIMATION AND SPECIFICATION
1. NOTATION AND UNITS

The terminology and notation used in this Item largely follow those employed by pump specialists.

Units

correction factor on total head rise for viscous liquids,

correction factor on volume flow rate for viscous liquids,

correction factor on efficiency for viscous liquids,

impeller diameter m

pipe internal diameter m

acceleration due to gravity, m/s2 m/s2

gauge total head referred to agreed vertical datum m

total head rise produced by pump, m

rotational speed of impeller rev/min

net positive suction head available referred to agreed vertical datum m

net positive suction head required referred to pump vertical datum m

specific speed, rad

pump shaft power W

absolute total pressure referred to agreed vertical datum Pa or N/m2

absolute static pressure Pa or N/m2

vapour pressure of pumped liquid Pa or N/m2

total pressure rise produced by pump, Pa or N/m2

volume flow rate through pump m3/s

suction specific speed, rad

vertical height, positive upwards measured from agreed vertical datum m

pump efficiency expressed as percentage, i.e. per cent

liquid dynamic viscosity N s/m2

CH CH H H∆ ′⁄∆=

CQ CQ Q Q′⁄=

Cη Cη η η′⁄=

D

d

g g 9.81≅

H

H∆ Hi∆ Ho Hi–=

N

NPSHa

NPSHr

nω nω ωQbep½ g Hbep∆( )¾⁄=

P

p

ps

pv

p∆ p∆ po pi–=

Q

Sω Sω ωQbep½ g NPSHrbep( )¾=

z

η η 100 Qρg H P⁄∆×=

µ

1

Issued November 1980 - 41 pages With Amendment A

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Subscripts

Superscript

liquid kinematic viscosity m2/s

density of pumped liquid kg/m3

angular speed of impeller rad/s

denotes atmospheric value

denotes value at pump best efficiency point

denotes value at pump inlet (suction) referred to pump datum elevation where relevan

denotes value at pump outlet (discharge) referred to pump datum elevation where relevant

prime denotes value associated with water as pumped liquid

ν ν µ /ρ=( )

ρ

ω

a

bep

i

o

′

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2. INTRODUCTION

This Item presents a method for estimating the size*, efficiency and typical flange sizes of a pump suitto a given duty. It may be used with the other Items in the series that give guidance to the non-specthe selection of pumps.

The other Items in the series are as follows.

2.1 Item Layout

In conjunction with Item No. 80030, Section 3 of this Item provides guidance on the determination ofappropriate design duty for approximate sizing. Section 4 then leads the user through a step-by-sprocedure which (1) indicates how pumps available from manufacturers' standard ranges may be ato meet the design duty and (2) provides estimates of the size of a suitable pump, its efficiency aof its suction and discharge flanges. A worked example in Section 5 is used to show the way in which thiinformation may be appraised in order to produce a specification for submission to a supplier. illustrates the dangers of specifying a pump to match an overestimated system resistance.

* The method provides an estimate of impeller diameter which is often used by manufacturers as a designation of the pump siz

Radial, mixed and axial-flow pumps. Introduction, Item No. 80030

Radial, mixed and axial-flow pumps. Glossary of terms, Item No. 81001

Radial, mixed and axial-flow pumps. Conversion factors, Item No. 81002.

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3. DESIGN DUTY

In the absence of other information, it may be assumed initially that the duty is to be covered by apump. The procedure in Section 4.1.1 indicates if this is feasible and, if not, gives guidance on how pumavailable from manufacturers' standard ranges may be arranged to cover the duty. With multarrangements, it should be noted that in order to cover contingencies such as the failure of one pudesign duty should be determined after consideration of the additional load on the pumps remaservice. Section 7.3 of Item No. 80030 (Reference 4) gives guidance.

In general, the design duty should be that which poses the most exacting pumping requireConsideration should be given to the following factors on which additional guidance is available inNo. 800304 as indicated.

A pump should be sized initially on the basis of Factor 1 and checked iteratively against Factor 2. 3 to 5 are seldom governing but, where they are, the penalties, e.g. of oversizing or efficiency, that may beincurred at the normal operating condition can be minimised; this may involve the use of variable spother controls on which guidance is given in Section 7.8 of Item No. 80030.

* In a system with changing liquid density, the maximum power requirement corresponds to the maximum liquid density. Note thapumpsized to produce a given total pressure rise under conditions of minimum liquid density (c.f. Factor 1) will produce the same total hearise but higher total pressure rise with the maximum liquid density. Thus a higher power will be absorbed.

Factor Considerations for Determining Design DutySection Numbers of

Item No. 800304

1. Full load flow rate at lowest temperature with largest likely systemlosses (determines most exacting and ). If a particular pressurerise requirement has to be met in a system with changing liquiddensity, should be calculated assuming the minimum value fordensity.

3.2, 6

2. Pump sized on the basis of Factor 1 but running at the highesttemperature with the smallest likely system losses (resulting increase in

determines most exacting and, usually * , see example inSection 5.4).

3.2, 3.4, 6, 7.2

3. Minimum load flow rate at highest temperature (determines mostexacting for high specific speed pumps).

3.4, 6

4. Minimum load flow rate at lowest temperature (determines mostexacting * for high specific speed pumps).

3.2

5. Shut-off head requirement for starting system (determines minimumshut-off head requirement).

3.2

Q H∆

H∆

Q NPSHr P

NPSHr

P

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4. SIZING AND SPECIFICATION

4.1 Size Estimation

The data presented in this section for estimating the size of a pump impeller and for estimating thethe connecting pipework are based on a correlation of manufacturers' data but, because designs diffethe data are approximate only. The procedure is geared towards the selection of the minimum size that will fulfil the design duty. For small scale installations, minimum size is the most usual selecriterion because it usually corresponds to the least capital cost. For larger installations, effconsiderations may be more important and, although the criteria of minimum size and maximum effare not fundamentally conflicting, the procedure may be readily adapted to provide an optimum effisolution.

4.1.1 Procedure

The following procedure leads to answers that are appropriate to single-stage single-inlet impelleleft to the discretion of the user to adapt the results to other configurations.

To use the procedure for a double-suction pump, work throughout with a value for that is half theflow rate. To use the procedure for a multi-stage pump, work with a total head rise, , that is thetotal head rise for the stage under consideration*.

The procedure requires values for the following variables:

* In most multi-stage pumps, the stages share the total head rise equally.

Liquid Properties at Design Point

(values for several liquids are available from Reference 2),

(values for several liquids are available from Reference 2),

(values for water are available from Reference 2).

System Properties at Design Point

(guidance on the calculation of this is given in Section 6 of Item No. 800304),

(per stage),

(per impeller inlet).

Step 1 Calculate from

.

Section 6 of Item No. 80030 gives guidance.

Step 2 If the liquid is water or is of lower viscosity than water, take , and jump to Step 4.

QH∆

pv

ν

ρ

Hi

H∆

Q

NPSHai

NPSHai Hi= pa pv–( ) ρg( )⁄+

Q′ Q= H ′ H∆=∆Cη 1=

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* Section 3.4 of Item No. 80030 gives an explanation. Note that increments in other than 0.5 m may be necesome cases.

If the pumped liquid is of higher viscosity than water, obtain a value for frFigure 1 using and . If , it may be worth considering a positidisplacement pump on efficiency grounds. Nevertheless, if only small powersinvolved, the user may be prepared to tolerate large losses of efficiency. Using theof obtained from Figure 1, calculate the ratio and obtain values for

and from Figure 2.

Step 3 Calculate the equivalent duty point for water as the pumped liquid, i.e.,

and .

Step 4 Decide on a suitable value for suction specific speed (e.g. rad for an ordinaryimpeller or rad for a special design). Allowing for the advisory 0.5 increment* on , take m and calculate , thmaximum speed before cavitation takes effect, from

.

Note , where is in rad/s and in rev/min.

Step 5 Check or against the available motor pole speeds using the following if direct-drive units are to be used. Select the highest speed that is below

.

If the available speeds are all too high, then (and ) should be increasedmay be achieved by one of the following means.

νmaxQ H∆ ν νmax≥

νmax ν νmax⁄ CQCH Cη

Q′ Q CQ⁄= H′∆ H CH⁄∆=

Sω 2.8= Sω 4=

NPSHai NPSHr NPSHai 0.5–= ωmax

NPSHai

ωmax Sω g NPSHr( )¾ /Q½=

Nmax30π------ ωmax= ω N

ωmax Nmaxωmax

Nmax

Nominal speeds of induction motors

Supply Frequency

No. of Poles 50 Hz 60 Hz

(rev/min) (rad/s) (rev/min) (rad/s)

2 2 900 305 3 500 366

4 1 450 152 1 750 183

6 960 100 1 160 121

8 720 75 870 91

10 570 60 690 72

N ω N ω

ωmax Nmax

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(i) Use a double-suction pump (halves ). Go back to Step 1.

(ii) Use more than one pump in parallel (reduces ). Go back to Step 1.

(iii) Fit an inducer (increases ) but see Section 3.4 of Item No. 80030.back to Step 4.

(iv) Increase the pump inlet pressure (increases ), e.g. by loweringthe pump. Go back to Step 1.

Step 6 Calculate the specific speed for the duty from

rad.

The most efficient pumps available have specific speeds between approximately 02 as illustrated by Sketch 3.3 of Item No. 800304. Thus, if , jump to Step10. If efficiency is not a prime consideration, proceed to Step 7. Otherwise, if a highervalue for is required, jump to Step 8, or, if a lower value is required, jump to Step 9.

Step 7 If , the specific speed falls within the range for standard pumps. JumStep 10.

If , proceed to Step 8.

If , jump to Step 9.

Step 8 Because is too low (see Section 3.7 of Item No. 80030 for details) a high speemulti-stage pump could be considered. If a high speed pump is preferred, go baStep 6 and choose a higher speed that is nevertheless below or . Howespeeds higher than 3600 rev/min are being considered, the manufacturer shoconsulted. If a multi-stage pump is preferred, go back to Step 2 with the amended valuefor .

Step 9 Because is too high (see Section 3.7 of Item No. 80030 for details), one or mthe following options must be taken.

(i) Reduce speed. Go back to Step 6.

(ii) Use a double-suction pump (halves ). Go back to Step 1.

(iii) Use more than one pump in parallel (reduces ). Go back to Step 1.

Q′

Q′

Sω

NPSHai

nω′ ωQ′½ g H′∆( )¾⁄=

0.6 nω′ 2≤≤

nω′

0.15 nω' 2≤ ≤

nω' 0.15<

nω' 2>

nω'

ωmax Nmax

∆H′

nω'

Q′

Q′

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Step 10 For viscous liquids, i.e. those resulting in , viscosity affects performancsignificantly. This effect has only been quantified1 for pumps of specific speed less tha1. If therefore and , it is advisable to select a pump of lower spespeed, see Step 9, unless the pump supplier can provide guidance.

If or , there is no difficulty. Proceed to Step 11.

Step 11 A particular shape of pump characteristic may be required to fulfil the sysrequirements, e.g.because there is more than one duty point to consider. By referrinSketches 3.4 to 3.6 of Item No. 80030, the behaviour of the characteristic cachecked and, if it is found that a different specific speed is required, the guidance in Step 8 or 9 may be used. Otherwise, proceed to Step 12.

Step 12 If a direct-drive induction motor pole speed has been selected, locate the approsize estimation chart from Figures 3a to 3j. Read off the value of the impeller diamete

(in mm), and the water efficiency, (in per cent), according to (in m) and(in m3/h or m3/s depending on the scale used). Alternatively, if other speeds havechosen, the formulae quoted on the charts must be used to obtain size and effiestimates. If the duty point is not covered by the graphed area of the approestimation chart, it is probable that the size or type of pump required is outsidstandard range offered by most manufacturers. However, by experimenting with pin parallel (or double-suction pumps) or pumps in series (or multi-stage pumps) atrying different speeds, a practical solution can usually be obtained.

Step 13 An estimate of the power required per stage per impeller inlet may be obtained fro

,

where . Note that here and correspond to the actual flow rate andhead rise values respectively rather than the water equivalent values.

Step 14 An approximate indication of the size of the pump suction bore is given by Figure4. Acorresponding indication of the discharge bore is given by Figure 5. Note that the fullflow rate through the pump should be used with Figures 4 and 5.

Cη 0.9<

nω′ 1≥ Cη 0.9<

nω′ 1≤ Cη 0.9≥

D η′ ∆H′ Q′

P Q Hρg∆

η------------------- 100×=

η η′ Cη⁄= Q ∆H

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4.2 Mechanical Suitability

For clean liquids that pose no special handling problems, it is likely that the type and size of pumpsat using the procedure of Section 4.1 will be available from manufacturers' standard ranges. If this iscase, Section 4.3 may be consulted directly for guidance with specification preparation. For other cthe following check list should be used.

Check Comment

Fibres in suspension

A specially-designed non-clogging pump is required for handling liqucontaining stringy or fibrous material such as found in sewage. Such a pshould be provided with an impeller, either constructed to cut the fibres, or forwith only one or two vanes to allow the material to pass freely through resulting large flow passages. The provision of an opening in the casingcleaning should be considered when specifying such a pump and, if the puliquid is liable to give off gases, the provision of a venting facility in the casshould also be considered. The efficiency of fibre-handling pumps may be lothan predicted from the procedure in Section 4.1.

Solids in suspension

The solids' handling capacity of a pump is measured in terms of “maximum spsize”, this being the largest diameter sphere that can be passed through the Solids' handling pumps are constructed more robustly than their clean licounterparts and the radial and mixed-flow types may be designed partially-open impellers. They may exhibit lower efficiencies than predicted frthe procedure in Section 4.1 and their efficiencies may be further reduced as weincreases the running clearances.

Corrosive or abrasive liquids

Most pump manufacturers can supply pumps made from, or coated with, matfor use with moderately corrosive or abrasive liquids. However, because it mabe possible to trim the impellers of such pumps to meet a given conditioservice (see Section 3.6 of Item No. 800304), some means of changing the pumspeed, e.g.by changing the wheel diameters of a belt transmission, shouldconsidered when preparing the specification.

Toxic or very high pressure liquids

The availability of glandless pumps appears to be restricted to specific speedsof less than approximately unity although there is no engineering reason forestriction.

High density or high viscosity liquids

Check that the pump components, particularly the shaft and impeller, are abwithstand the additional torque required.

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4.3 Specification

When the type of pump or combination of pumps has been finally decided, a full specification mgiven to those manufacturers invited to tender. The form at the back of this Item is suitable fospecifications. The form should first be partly filled in by the purchaser to convey sufficient informto the manufacturer to enable a satisfactory quotation to be given, e.g. the duty point, preference forhorizontal or vertical orientation, etc. should be specified. Then the manufacturer can complete the or suggest alternative designs. The example given in Section 5.3 clarifies the use of the form and thGlossary of Pump Terms, Item No. 810015 covers its terminology.

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5. EXAMPLE

A pump is required to raise water at a constant temperature of C from a sump to an open reservalternative pump installations are proposed for meeting the design flow of 260 m3/h (0.0722 m3/s), seeSketch 5.1. In Installation A, at the design flow rate, the frictional loss is 0.05 m for the suction sy(stations 2 to 3) and 59.95 m for the discharge system (stations 4 to 5). In Installation B, at the desirate, the loss due to friction in the suction system is 0.5 m (stations 2 to 3) and 59.5 m (stations 4the discharge system. The supply and receiving reservoirs have the same plan area and thus the levat a difference of 10 m. Additional data are shown in Sketch 5.1.

For a 50 Hz electricity supply, investigate the sizes and power requirements of both installations anda specification data sheet for the preferred installation. Investigate the consequences of the frictionin the discharge system being overestimated by 12 m total head and discuss the possible courses (This part of the example illustrates the dangers of over-design and the application of the pump simlaws.)

Sketch 5.1 Alternative installations*

5.1 Fully-Submerged Pump (Installation A)

Following through the procedure in Section 4.1.1, a single-stage, single-inlet pump will initially bproposed. The flow rate and head rise requirement of the system thus remain unfactored.

The values of the necessary variables are as follows.

* A more practical pump datum may be the inlet bellmouth since, at a preliminary investigation stage such as this, the exact position of theimpeller inlet eye may not be known. However the difference is small and errs on the safe side for the calculation.

60°

NPSHa

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and the

anging

,

Liquid Properties

Pa at C from Item No. 680102,

is not required as the liquid is water,

kg/m3 at C from Item No. 680102.

System Properties

Section 6 of Item No. 800304 illustrates a method for calculating , the total head at the pump inlet refeto the pump datum elevation.

Taking the minimum sump water level as the system datum, i.e. , the total head, , at the puminlet referred to the system datum is given by

, c.f. Equation (6.10) of Item No. 80030,

where since the total lead loss in the sump is negligible and

m, which is the frictional loss in the suction system.

Thus, m.

The value of is then given by

, c.f. Equation (6.9) of Item No. 80030,

where , the head adjustment allowing for the difference in elevation between the system datum pump datum, is equal to m.

Thus, m.

Again, using Section 6 of Item No.80030 as a guide, the pump total head rise, , follows from rearrthe equation

,

c.f. Equation (6.4) of Item No. 80030, i.e.,

,

c.f. Equation (6.5) of Item No. 80030.

In this example, since , m, , and m

m.

pv 0.19225×10= 60°

ν

ρ 983.15= 60°

Hi

Hi 0= H3

H3 H1 H1 H2–( )– H2 H3–( )–=

H1 H2 0=–

H2 H3– 0.05=

H3 0 0– 0.05 0.05–=–=

Hi

Hi H3 δz–=

δz 3–

Hi 0.05– 3 2.95=+=

∆H

H1 H1 H2–( )– H2 H3–( )– H H4 H5–( )– H5–∆+ 0=

H∆ H5 H1– H1 H2–( ) H2 H3–( ) H4 H5–( )+ + +=

H1 0= H5 10= H1 H2–( ) 0= H2 H3–( ) 0.05= H4 H5–( ) 59.95=

H∆ 10 0– 0 0.05 59.95 70=+ + +=

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ps.

The design flow rate is m3/s.

Step 1 ,

where , the atmospheric pressure, may be taken as Pa. Thus,

m.

Step 2 The liquid is water so

m3/s,

m and

.

Jump to Step 4.

Step 4 Taking rad and m,

rad/s or rev/min.

Step 5 The highest speed available for a 50 Hz supply that is below is the 2 pole spe305 rad/s (2900 rev/min).

Step 6 The specific speed for the duty is given by

rad.

This is within the range for which high efficiency pumps are available so there ireason to change .

Step 7 Jump to Step 10 because the specific speed falls within the range for standard pum

Step 10 Proceed to Step 11 as there is no need to consider viscosity effects.

Step 11 No particular shape of pump characteristic is preferred.

Step 12 Figure 3a is the size estimation chart appropriate to rad/s (rev/min). From this with m and m3/s, the estimated impellerdiameter is 260 mm and the estimated efficiency is 81 per cent.

Step 13 W.

Q 0.0722=

NPSHai Hi pa pv–( ) ρg( )⁄+=

pa 1.0135×10

NPSHai 2.95 1.013 5×10 0.1992 5×10–( ) 983.15 9.81×( )⁄ 11.4=+=

Q′ 0.0722=

H′∆ 70=

Cη 1=

Sω 2.8= NPSHr NPSHai 0.5–=

ωmax 2.8 9.81 10.9×( )¾ 0.0722½

346.5=⁄= Nmax 3 309=

ωmax

nω′ 305 0.0722½ 9.81 70×( )¾⁄× 0.61==

nω′

ω 305= N 2900=H′ 70=∆ Q′ 0.0722=

P 0.0722 70 983.15 9.81×××

81------------------------------------------------------------------ 100 60

3×10=×=

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5.2 Above Ground Mounted Pump (Installation B)

Following through the procedure in Section 4.1.1, a single-stage, single-inlet pump will again initially bproposed. The flow rate and head requirements of the system thus remain unfactored.

The values of the necessary variables are as follows.

Liquid Properties

Pa at C from Item No. 680102,

is not required,

kg/m3 at C from Item No. 680102.

System Properties

Following the method used in Section 5.1,

, i.e.,

m and, with

, then

m.

, i.e.,

m.

m3/s.

Step 14 From Figure 4, a typical diameter for the suction bore would be 150 mm and, frFigure 5, a corresponding discharge bore would be 125 mm.

Step 1 From ,

m.

Step 2 m3/s,

m and

.

Jump to Step 4.

Step 4 Taking rad and m,

pv 0.19925×10= 60°

ν

ρ 983.15= 60°

H3 H1 H1 H2–( )–= H2 H3–( )–

H3 0 0.5 0.5–=–=

Hi H3 δz–=

Hi 0.5– 3.5– 4–==

H H5 H1– H1 H2–( ) H2 H3–( ) H4 H5–( )+ + +=∆

H 10 0– 0 0.5 59.5 70=+ + +=∆Q 0.0722=

NPSHai Hi pa pv–( ) ρg( )⁄+=

NPSHai 4– 1.0135×10 0.1992

5×10–( ) 983.15 9.81×( )⁄ 4.4=+=

Q′ 0.0722=

H′ 70=∆

Cη 1=

Sω 2.8= NPSHr NPSHai 0.5–=

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* Although the use of a two-stage pump is the most satisfactory alternative for the conditions of this example, the double-suction radial or mixed-flow pump should also be generally considered where higher motor speeds are availabgiven value of , the use of a double-suction pump increases by a factor of 2 which in many cases will allow amotor pole speed and thus a smaller impeller diameter to be selected.

rad/s or rev/min

Step 5 The highest speed available for a 50 Hz supply that is below is the 4 pole sp152 rad/s (1450 rev/min).

Step 6 The specific speed for the duty is given by

rad.

This is below the range in which high efficiency pumps can be obtained but as thno special efficiency optimisation requirement, the procedure is continued to Step 7.

Step 7 Jump to Step 10 because the specific speed falls within the range for standard pum

Step 10 Proceed to Step 11 as there is no need to consider viscosity effects.

Step 11 No particular shape of pump characteristic is preferred.

Step 12 Figure 3b is the size estimation chart appropriate to rad/s ( rev/mFrom this, with m and m3/s, the suggested impeller diameter

mm and is likely to be larger than those available from most manufactustandard ranges because the duty point is not covered by the graphed area of thHowever, a two-stage pump with each stage producing 35 m total head rise is a poalternative*. To estimate the size and power requirements of such a pump, the procis repeated from Step 6 onwards taking m3/s as before but m.

Step 6 rad.

Step 7 Jump to Step 10.

Step 10 Proceed to Step 11.

ωmax 2.8 9.81 3.9×( )¾ 0.0722½⁄ 160== Nmax 1530=

ωmax

nω′ 152 0.0722½ 9.81 70×( )⁄× ¾ 0.30==

ω 152= N 1450=H′ 70=∆ Q′ 0.0722=

450>

Sω Nmax

Q′ 0.0722= H′ 35=∆

nω′ 152 0.0722½× 9.81 35×( )¾ 0.51=⁄=

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5.3 Discussion

Although either installation is able to deliver the duty flow rate, Installation A is preferable not only beit requires a more simple single-stage pump than Installation B but primarily because of the pdifficulties that would accompany Installation B. Installation B, however, has the advantage of easierfor maintenance.

As indicated in Section 4.3, the detail to which the pump can be specified may vary according to whthe purchaser has a preference for (and adequate knowledge of) certain features or whether he or sto accept the recommendations of the pump manufacturer. Certain data, however, are essential asby the letter “P” in the annotated columns of the following data sheet which has been filled in with eappropriate to the example. The lines in the form are annotated according to the following legend:

Step 11 Proceed to Step 12.

Step 12 From Figure 3b again, but this time with m3/s and m, theestimated impeller diameter is 360 mm and the estimated efficiency is 80 per cent

Step 13 W per stage or W for thepump.

Step 14 From Figure 4, a typical diameter for the suction bore would be 200 mm and, frFigure 5, a corresponding discharge bore would be 150 mm.

P entry from purchaser expected.

P/V according to responsibility, e.g. for supplying driving unit, entry is expected from eithpurchaser or vendor.

O entry from purchaser optional.

OV entry from purchaser optional. If left blank, entry from vendor is expected.

V entry from vendor expected.

Q′ 0.0722= H′ 35=∆

P 0.0722 35 983.15 9.81×××

80------------------------------------------------------------------ 100 30.5

3×10=×= 613×10

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Sketch 5.2 Critical installation dimensions

5.4 Effects of Overestimated Discharge System Total Head Loss

5.4.1 Equilibrium flow rate

In order to investigate the effects of the estimated system loss being in error, it is first necessary toflow rate at which the pump and actual system are in equilibrium. As explained in Section 7.1 of Ite800304, this may be found by plotting the system total head requirement characteristic on the samas the pump total head rise characteristic and noting the flow rate at which the two curves intersec

For plotting the system total head requirement characteristic, it is sufficiently accurate to assume frictional component of the total head requirement is directly proportional to the flow rate squaredapproach has been used for constructing the system curve illustrated in Sketch 5.3.

The pump total head rise characteristic may in this example be taken at constant speed for, althopower required is different from the design value, the electric motor driving unit will maintain the sclose to the nominal 2-pole speed of 2900 rev/min. The constant speed pump total head rise charafor 2900 rev/min should preferably be obtained from the manufacturer but, in the absence oinformation, may be estimated using Sketch 3.4 of Item No. 80030. In this example, manufacturehave been used to construct the pump curve illustrated in Sketch 5.3.

Sketch 5.3 shows that the actual system characteristic is in equilibrium with the pump at a flow rate,0.0774 m3/s with a total head rise, , of 65 m. The increase in flow rate above the required va0.0722 m3/s is thus fairly small.

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Sketch 5.3 Change of operating point resulting from overestimated system losses

5.4.2 Pump and system properties at equilibrium flow rate

The power required, efficiency and may now be obtained from the manufacturer's curestimated using Sketches 3.5, 3.6 and 3.11 respectively in Section 3 of Item No. 800304 with

.

Preferably, the manufacturer's data should be used from which:

W,

per cent,

m.

For the system running at m3/s, may be estimated assuming that the frictioncomponent of total head loss in the suction system is directly proportional to the flow rate squa

m3/s, this frictional component was 0.05 m, so at m3/s it is

m.

Thus the corresponding value for is given by

m.

This value for , although only slightly lower than that at the required flow rate, is less thanew value for . Therefore, if the pump is to run continuously under such conditions, cavitatio

NPSHr

Q Q⁄ bep 0.0774 0.0722⁄ 1.07==

P 60.63×10=

η 80=

NPSHr 11.4=

Q 0.0774= NPSHai

Q 0.0772= Q 0.0774=

0.05 0.0774 0.0722⁄( )× 20.0575=

NPSHai

NPSHai( )Q 0.0774=

NPSHai( )Q 0.0722=

0.0575 0.05–( )–=

11.4 0.0075 11.39=–=

NPSHaiNPSHr

19

80031�rgin of

stic

ad

sary. In

1039rifice to

o avoid

ty laws

y frome lostment.equate of the

cause considerable damage. Note that although and are not very different, a maat least 0.5 m is normally advised, see Section 3.4.2 of Item No. 800304.

5.4.3 Courses of action

Given that it is not possible to increase , one of three courses of action may be taken.

(1) Throttle the pump discharge (e.g. by installing an orifice plate) to restore the system characterito the original estimate.

(2) Reduce the speed of the pump (e.g. by installing a belt transmission) to lower the pump total herise curve so that it intersects with the actual system curve at the required flow rate of m3/s.

(3) Reduce the diameter of the impeller.

(1) Pump discharge throttling

This is the simplest solution but it results in the pump absorbing more power than is necesorder to restore the system total head loss of 58 m at a flow rate of 0.0722 m3/s back to the estimatedvalue of 70 m, an orifice plate is required to cause a total head loss of 12 m. Item No. 83

provides pressure loss data on several orifice geometries and enables the correct size of obe calculated. Note that orifice plates should be installed on the discharge side of pumps treductions in .

(2) Speed reduction

Reducing the speed lowers the pump total head rise characteristic according to the similarigiven in Section 3.5 of Item No. 800304.

Provided the actual system curve intersects the modified pump characteristic no further awathe best efficiency point than the original pump characteristic, no pumping efficiency need band a saving in power will result from the reduction in the system total head rise requireAlthough the speed reduction calculation is, for generality, an iterative process, an adapproximation to the required speed can, in many cases, be achieved within the first stepiteration.

NPSHai NPSHr

NPSHai

Q 0.0722=

NPSHai

20

80031�

throughw aion (on (oniredulated

tested.peedse

Sketch 5.4 Shift of operating point due to change of speed

The object of the calculation is to find the speed that causes the pump characteristic to passthe required operating point, see Sketch 5.4. Because the system characteristic does not follospeed transformation path (square law), the required operating point is at a different posittransformation path B) on the pump characteristic from that of the actual operating pointtransformation path A). It is this feature that prevents the similarity laws from relating the requspeed directly to the actual and required operating points. The calculation is therefore formas follows.

For the purpose of starting the iterative calculation, a new speed must be postulated andBecause pathA does not intersect with the required operating point, an average of the two scorresponding to (1) the required head change down path A and (2) the required flow rate changdown path A is taken for the initial estimate.

21

80031�

1) androduce.0722

tersects

may

Sketch 5.5 Effect of speed reduction

Because each speed calculation follows a speed transformation path, i.e. moves to the same relativeposition on the pump characteristics, the similarity laws are applicable. Thus Equations (3.1(3.12) of Item No. 800304 may be used and rearranged to estimate the speeds required (1) to pthe desired total head rise, , of 58 m and (2) to produce the required flow rate, , of 0m3/s. Thus, from Equation (3.11) of Item No. 800304,

rev/min,

whereas, from Equation (3.12) of Item No. 80030,

rev/min.

Averaging these speeds gives rev/min.

The next stage is to plot the pump characteristic corresponding to this new speed to see if it inwith the system characteristic acceptably close to the required flow rate of 0.0722 m3/s.

Equations (3.11) and (3.12) of Item No. 800304 are used to transform points A and bep on curve 1to their same relative positions on curve 2 (for rev/min) so that a portion of curve 2be constructed.

Thus,

m,

m3/s,

∆H Q

N2 N1

H2∆H1∆

---------- ½

29005865------

½2739===

N2 N1Q2 Q1⁄ 2900 0.0722 0.0774 2705 =⁄×==

N2 2722=

N 2722=

H2A H1A

N2

N1------

∆=∆2

6527222900------------

257.27==

Q2A Q1AN2 N1 0.0774 2722 2900⁄ ×=⁄ 0.07265==

22

80031�

ng to

me speed

r error in

using.11 of

eristic

aws

tween

otor but pump

m and

m3/s.

By joining points and together, a portion of the pump characteristic correspondirev/min is formed. It can be seen from Sketch 5.5 that this intersects with the system

characteristic, i.e. to form a new operating point, 2B, at m3/s which is acceptably closeto the required flow of 0.0722 m3/s. Note that if this error were not acceptable, and in socircumstances it may be much larger, the calculation could be repeated by basing the newon the average of those required to correct the lesser error in total head rise and the lesseflow rate.

Having found the required speed, a check must be made on the power required and the manufacturer's data that are available for rev/min only (or Sketches 3.5 and 3Item No. 80030 if such data are not available). At the new operating point,

.

Then, using this ratio to find the equivalent flow rate, , for rev/min,

m3/s.

With this equivalent flow rate, the values for power and corresponding to point 1B mayalso be found from the manufacturer's data, i.e.

W and

m.

These values, although applicable to rev/min, lie on the same part of the charactcurve (i.e. speed transformation path B) as the required operating point, 2B. They may thereforebe directly transformed into values applicable to rev/min using the similarity lgiven in Section 3.5 of Item No. 80030 as follows.

W,

m.

Note that is unchanged and thus remains at 11.4 m, leaving a margin of 1.53 m be and . A power saving of W results.

The required speed of 2722 rev/min does not correspond to the pole speed of an electric mit can nevertheless be achieved economically through a belt transmission for which mostmanufacturers are able to quote.

H2bep H1bep

N2

N1------

2

7027222900------------

261.67==∆=∆

Q2bep Q1bepN2 N1⁄ 0.0722 2722 2900⁄ × 0.06777===

2bep 2AN2 2722=

Q 0.0718=

NPSHrN 2900=

Q2B Q2bep⁄ 0.0718 0.06777⁄ 1.059= =

Q1B N 2900=

Q1B Q1bepQ2B Q2bep⁄ 0.0722 1.059× 0.07646= ==

NPSHr

P1B 60.63×10=

NPSHr1B 11.2=

N 2900=

N2 2722=

P2B P1B27222900------------

350.1

3×10==

NPSHr2B NPSHr1B27222900------------

29.87==

NPSHaiNPSHai NPSHr2B P1A P2B– 10.5 3×10=

23

80031�

thess bepeller

powerpeed

d fromt to a

out oflate is

e wheelhas been

(3) Impeller size reduction

Section 3.6 of Item No. 80030 provides information on impeller trimming which, for requirements given in this example, is likely to be a practical measure. It will neverthelenecessary to seek the advice of the pump manufacturer before undertaking to trim the imbecause the performance of an impeller can be very sensitive to details of geometry. Thesaving following impeller trimming is likely to be close to that resulting from changing the sto suit the system characteristic, i.e. approximately W.

5.4.4 Best course of action

The following factors should be considered:

power savings resulting from modification,

cost and ease of modification,

ability of modified system or pump to adapt to changing conditions.

Power savings

No saving of power results from installing an orifice whereas a saving of kW can be expectereducing the speed or trimming the impeller. For a pump running continuously, this can amounsubstantial saving over a period of years.

Cost and ease of modifications

The installation of a belt transmission or the trimming of the impeller will cause the pump to be service for a considerably longer period than the installation of an orifice. The cost of an orifice pminimal in relation to the alternative modifications.

Adaptability

If the system or pump characteristic changes, the orifice plate can be changed to suit. Similarly, thsizes on the belt transmission may also be changed to vary the speed to suit. Once the impeller trimmed, however, only decreases in the system total head requirement can be accommodated.

10 3×10

10∼

24

80031�

listed in

in the

3th

d 5.

ing

o.

No.

em

.

6. REFERENCES AND SOURCES OF DATA

6.1 References

The references given are recommended sources of information supplementary to that in this Item (chronological order)

6.2 Sources of Data

The following establishments and manufacturers supplied data or other information employedproduction of this Item (listed in alphabetical order).

1. – Standards for centrifugal, rotary and reciprocating pumps, 1edn, Hydraulic Inst., Cleveland, Ohio, 1975.

2. ESDU Physical data, chemical engineering sub-series, Vols 1, 3 anEngineering Sciences Data Unit, 1980.

3. ESDU Fluid mechanics, internal flow sub-series. Vols 1-4. EngineerSciences Data Unit, 1980.

4. ESDU Radial, mixed and axial-flow pumps. Introduction. Item N80030, Engineering Sciences Data Unit, 1980.

5. ESDU Radial, mixed and axial-flow pumps. Glossary of terms. Item 81001, Engineering Sciences Data Unit, 1981.

6. ESDU Radial, mixed and axial-flow pumps. Conversion factors. ItNo. 81002, Engineering Sciences Data Unit, 1981.

7. DURCO Durion Co. Inc. Pump Division, Dayton, Ohio 45401, USA.

8. GEC GEC Mechanical Handling Ltd, Erith, England.

9. GIRDLESTONE Girdlestone Pumps Ltd, Woodbridge, Suffolk, England.

10. HAYWARD TYLER Hayward Tyler Ltd, Luton, England.

11. PEERLESS Peerless Pump, Montebello, California 90640, USA.

12. SALMSON Société Electro-Hydraulique, 92213 St Cloud, Cedex, France

13. SCANPUMP Scanpump AB, Mölndal, Sweden.

14. SPP Sigmund Pulsometer Pumps Ltd, Reading, England.

15. STORK J. & S.-Stork Pumps Ltd, Horley, Surrey, England.

16. WEIR Weir Pumps Ltd, Glasgow, Scotland.

17. WORTHINGTON-SIMPSON Worthington-Simpson Ltd, Newark, England.

25

80031�

26

80031�

FIGURE 1 MAXIMUM KINEMATIC VISCOSITY, υmax

Flow rate per impeller inlet, Q (m3/s)

2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101

Head riseper stage,∆H (m)

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104

60

40

30

20

2 + 10 -5

3 + 10 -5

4 + 10 -5

6 + 10 -5

8 + 10 -510 -4

1.5 + 10 -42 + 10 -4

3 + 10 -4

4 + 10 -4

6 + 10 -4

8 + 10 -410 -3

1.5 + 10 -3

2 + 10 -3

3 + 10 -3

4 + 10 -3

6 + 10 -3

vmax (m2/s)

vmax (cSt)

Flow rate per impeller inlet, Q (m3/h)

4 8 2 5 8 2 5 8 2 5 8 2 3

1

1

80 100 150 200 300 400 600 800 1000 1500 2000 3000 4000 6000

27

80031�

FIGURE 2 CORRECTION FACTORS CH , CQ , Cη FOR VISCOUS LIQUIDS

CH

0.7

0.8

0.9

1

CQ

0.7

0.8

0.9

1

Cη

0.5

0.6

0.7

0.8

0.9

1

(Q)(∆Hg)¼/ν

2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8103 104 105 106

ν / νmax

2

3

4

5678

2

3

4

5678

2

0.01

0.1

1

NOTE. Obtainνmax from Figure 1,

calculate ν / νmax for

the pumped liquid andread off values forCH, CQ and Cη from

this Figure by followingthe arrows.

Curves derived from Reference 1

�

�

�

�

�

28

80031�

FIGURE 3a SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 2900 rev/min (50 Hz)

Water equivalent flow rate per impeller inlet, Q1 (m3/s)

2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101

Waterequivalenthead riseper stage,∆H1

(m)

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104

ImpellerdiameterD (mm)

Water equivalent efficiency, η' (per cent)

2900 rev/min

450400350

300

250

200

175

150

30

4050

60 65 70 72.5 75 77.5

80

82.5

85

87.5

125

Water equivalent flow rate per impeller inlet, (m3/h)

4 8 2 4 6 8 2 4 6 8 2 4 6 8 2 3

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity

Equation for D valid for 0.15 � n'ω � 2,

Equation for η' valid for 0.15 � n'ω � 6 with Q'min � Q' � Q'max

where Q'min is the greater of 1+10-3 m3/s or 0.02 n1ω2 m3/s

and Q'max = 0.83 n1ω2 m3/s

D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

29

80031�

FIGURE 3b SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 1450 rev/min (50 Hz)

Water equivalent flow rate per impeller inlet, Q' (m3/s)

2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101

Waterequivalenthead riseper stage,∆H'(m)

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104

Impellerdiameter D (mm)


1450 rev/min

450

400

350

300

250

200

175

15030

4050 60 65 70 72.5 75 77.5

80

82.5

85

87.5


4 8 5 82 5 82 5 82 32

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

30

80031�

FIGURE 3c SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 960 rev/min (50 Hz)


2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101


2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103



960 rev/min

450

400

350

300

250

200

17530

4050

60 65 70 72.5 7577.5

80

82.5

85

87.5


4 8 5 82 5 82 5 82 3210410310210

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

31

80031�

FIGURE 3d SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 720 rev/min (50 Hz)


2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101


2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103



720 rev/min

450

400

350

300

250

20030

4050

60 65 70 72.5 7577.5

80

82.5

85


4 8 5 82 5 82 5 82 32

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

10410310210

32

80031�

FIGURE 3e SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 570 rev/min (50 Hz)


2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101


2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104



570 rev/min

450

400350

300

250

200

30

4050

60 65 70 72.5 75 77.5

82.5

85


4 8 322 4 6 82 4 6 8 2 4 6 8

80

Derivation

=

+

2 7 1 822

0 26

2. .

.n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

33

80031�

FIGURE 3f SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 3500 rev/min (60 Hz)


2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101


2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104



3500 rev/min

450400350

300

250

200

175

150

30

4050

60 65 70 72.5 7577.5

80

82.5

85

87.5

125


4 8 322 4 6 8 2 4 6 8 2 4 6 8

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

34

80031�

FIGURE 3g SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 1750 rev/min (60 Hz)


2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101


2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104



1750 rev/min

450

400

350

300

250

200

175

15030

4050

60 65 70 72.5 75 77.5

80

82.5

85

87.5


4 8 322 4 6 82 4 6 8 2 4 6 8

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

35

80031�

FIGURE 3h SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 1160 rev/min (60 Hz)


2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101


2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104



1160 rev/min

450

400

350

300

250

200

175

30

40 5060 65 70 72.5 75

77.5

80

82.5

85

87.5


4 8 322 4 6 82 4 6 8 2 4 6 8

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

36

80031�

FIGURE 3i SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 870 rev/min (60 Hz)


2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101


2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104



870 rev/min

450

400

350

300

250

200

17530

4050 60 65 70 72.5 75

77.5

80

82.5

85


4 8 322 4 6 82 4 6 8 2 4 6 8

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

37

80031�

FIGURE 3j SIZE AND EFFICIENCY ESTIMATION CHART FOR N = 690 rev/min (60 Hz)


2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1 100 101


2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

2

3

4

5

678

10-1

100

101

102

103101 102 103 104



690 rev/min

450400

350

300

250

20030

4050

60 65 70 72.5 7577.5

82

85


4 8 322 4 6 82 4 6 8 2 4 6 8

80

Derivation

=

+

2 7 1822

0 26

2. ..n n' 'ω ω

Validity





D d n H g d= ' ' ( ' )ω ω ωω∆1

2 (m) where '

η1 = 96 - {23.26 Q1-0.43 n1ω

-1.36 + 43.16 Q1-0.4n1ω} ½ per cent.

38

80031�

S

2 3 4 5 6 7 8100

argest likely diameter

Smallest likely diameter

39

FIGURE 4 APPROXIMATE CORRELATION OF SUCTION BORE DIAMETER

Q 2900/N (m3/s) / ( rev/min )

2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1

Di

(mm)

2

3

4

5

6

78

2

3

4

5

6

78

101

102

103

Standard size to BS 5257 (ISO 2858)for horizontal end-suction centrifugal pumpsSee Section 5, Item No 80030 L

40

80031�

RS

2 3 4 5 6 7 8100

Largest likely diameter

Smallest likely diameter

FIGURE 5 APPROXIMATE CORRELATION OF DISCHARGE BORE DIAMETE

Q 2900/N (m3/s) / (rev/min )

2 3 4 5 6 7 8 2 3 4 5 6 7 810-3 10-2 10-1

D0

(mm)

2

3

4

5

6

7

8

2

3

4

5

6

7

8

101

102

103

Standard size to BS 5257 (ISO 2858)for horizontal end-suction centrifugal pumpsSee Section 5, Item No 80030

∆H > 32 (N/2900)2 m(rev/min)2

∆H � 32 (N/2900)2

m(rev/min)2

80031�

ar Item

eeringhanicalpointed

gineers.

eeringinitialItem was

THE DEVELOPMENT OF THIS DATA ITEM

The work of the permanent professional staff of the Engineering Sciences Data Unit on this particulwas monitored and guided by the following Working Party:

Mr D. Burgoyne – J&S-Stork Pumps LtdMr T. Cuerel – B.P. Trading LtdMr D.J. Luget – Kellogg International CorporationMr D.S. Miller – British Hydromechanics Research AssociationMr P.H. Nuttall – Sigmund Pulsometer Pumps LtdDr I. Pearsall – National Engineering LaboratoryDr D. Pollard – GEC Power Engineering Ltd, WhetstoneMr D.W. Standish – Girdlestone Pumps Ltd,

on behalf of the Internal Flow Panel which has the following constitution:

ChairmanMr N.G. Worley – Babcock Power Ltd

MembersMr J. Campbell – Ove Arup PartnershipDr D. Chisholm – National Engineering LaboratoryDr D.J. Cockrell – Leicester UniversityDr R.B. Dean – Atkins Research and DevelopmentMr D.H. Freeston* – Auckland University, New ZealandDr G. Hobson – GEC Turbine Generators Ltd, RugbyProf. J.L. Livesey – Salford UniversityMr D.S. Miller – British Hydromechanics Research AssociationMr B. Payne – Kellogg International CorporationDr D. Pollard – GEC Power Engineering Ltd, WhetstoneMr J.A. Ward – Atomic Energy Technology Unit.

The Internal Flow Panel has benefited from the participation of members from several engindisciplines. In particular, Dr G. Hobson has been appointed to represent the interests of mecengineering as the nominee of the Institution of Mechanical Engineers and Mr B. Payne has been apto represent the interests of chemical engineering as the nominee of the Institution of Chemical En

The work on this Item was carried out in the Internal Flow and Physical Properties Group of the EnginSciences Data Unit. The member of staff who undertook the technical work involved in the assessment of the available information and the construction and subsequent development of the

Mr C.J.T. Clarke – Group Head, Internal Flow and Physical Properties.

* Corresponding Member

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Documents

Pumps Specification 80031E