How Hydraulic Pumps and Motors work

how electric hydraulic pumps work and how to calculate hydraulic pump flow rate and how to size hydraulic pumps and motors how to wire hydraulic pumps
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Dr.NaveenBansal,India,Teacher
Published Date:25-10-2017
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2 HydraulicPumps 2.1 INTRODUCTION Hydraulics is defined as the science of the conveyance of liquids through pipes. Thepumpisoftenusedtoraisewaterfromalowleveltoahighlevelwhereitcan be stored in a tank. Most of the theory applicable to hydraulic pumps has been derivedusingwaterastheworkingfluid,butotherliquidscanalsobeused.Inthis chapter, we will assume that liquids are totally incompressible unless otherwise specified. This means that the density of liquids will be considered constant no matterhowmuchpressureisapplied.Unlessthechangeinpressureinaparticular situation is very great, this assumption will not cause a significant error in calculations. Centrifugal and axial flow pumps are very common hydraulic pumps. Both work on the principle that the energy of the liquid is increased by imparting kinetic energy to it as it flows through the pump. This energy is suppliedbytheimpeller,whichisdrivenbyanelectricmotororsomeotherdrive. The centrifugal and axial flow pumps will be discussed separately in the following sections. 2.2 CENTRIFUGALPUMPS Thethreeimportantpartsofcentrifugalpumpsare(1)theimpeller,(2)thevolute casing, and (3) the diffuser. Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved48 Chapter2 2.2.1 Impeller The centrifugal pump is used to raise liquids from a lower to a higher level by creating the required pressure with the help of centrifugal action. Whirling motionisimpartedtotheliquidbymeansofbackwardcurvedbladesmountedon awheelknownastheimpeller.Astheimpellerrotates,thefluidthatisdrawninto thebladepassagesattheimpellerinletoreyeisacceleratedasitisforcedradially outwards.Inthisway,thestaticpressureattheouterradiusismuchhigherthanat theeyeinletradius.Thewatercomingoutoftheimpelleristhenleadthroughthe pump casing under high pressure. The fluid has a very high velocity at the outer radius of the impeller, and, to recover this kinetic energy by changing it into pressure energy, diffuser blades mounted on a diffuser ring may be used. The stationary blade passages have an increasing cross-sectional area. As the fluid movesthroughthem,diffusionactiontakesplaceandhencethekineticenergyis convertedintopressureenergy.Vanelessdiffuserpassagesmayalsobeused.The fluid moves from the diffuser blades into the volute casing. The functions of a volute casing can be summarized as follows: It collects water and conveys it to the pump outlet. The shape of the casing is such that its area of cross-section gradually increases towards the outlet of the pump. As the flowing water progresses towards the delivery pipe, more and more water is added from the outlet periphery of the impeller. Figure 2.1 shows a centrifugal pump impeller with the velocity triangles at inlet and outlet. For the best efficiency of the pump, it is assumed that water enters the impeller radially, i.e.,a ¼ 908 and C ¼ 0. Using Euler’s pump equation, the 1 w1 work done per second on the water per unit mass of fluid flowing W E ¼ ¼ðÞ U C 2U C ð2:1Þ 2 w2 1 w1 m Where C is the component of absolute velocity in the tangential direction. E is w referred to as the Euler head and represents the ideal or theoretical head developed by the impeller only. The flow rate is Q ¼ 2pr C b ¼ 2pr C b ð2:2Þ 1 r1 1 2 r2 2 WhereC istheradialcomponentofabsolutevelocityandisperpendiculartothe r tangentattheinletandoutletand bisthewidthoftheblade.Forshocklessentry and exit to thevanes,water enters and leaves the vane tips in adirection parallel to their relative velocities at the two tips. AsdiscussedinChapter1,theworkdoneonthewaterbythepumpconsists of the following three parts: 2 2 1. The part (C – C )/2 represents the change in kinetic energy of the 2 1 liquid. 2 2 2. The part (U – U )/2 represents the effect of the centrifugal head or 2 1 energy produced by the impeller. Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 49 Figure 2.1 Velocity triangles for centrifugal pump impeller. 2 2 3. The part (V 2 V )/2 represents the change in static pressure of the 2 1 liquid, if the losses in the impeller are neglected. 2.3 SLIPFACTOR From the preceding section, it may be seen that there is no assurance that the actual fluid will follow the blade shape and leave the impeller in a radial direction. There is usually a slight slippage of the fluid with respect to the blade rotation. Figure 2.2 shows the velocity triangles at impeller tip. 0 InFig.2.2,b istheangleatwhichthefluidleaves theimpeller,andb is 2 2 0 the actual blade angle, and C and C are the tangential components of w2 w2 0 absolutevelocitycorrespondingtotheanglesb andb ,respectively.Thus,C 2 2 w2 0 is reduced to C and the difference DC is defined as the slip. The slip factor w2 w is defined as 0 C w2 Slip factor;s ¼ C w2 According to Stodola’s theory, slip in centrifugal pumps and compressors is due torelativerotationoffluidinadirectionoppositetothatofimpellerwiththesame Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved50 Chapter2 Figure 2.2 Velocity triangle at impeller outlet with slip. angular velocity as that of an impeller. Figure 2.3 shows the leading side of a blade,wherethereisahigh-pressureregionwhileonthetrailingsideoftheblade there is a low-pressure region. Duetothelowerpressureonthetrailingface,therewillbeahighervelocity andavelocitygradientacrossthepassage.Thispressuredistributionisassociated withtheexistenceofcirculationaroundtheblade,sothatlowvelocityonthehigh- pressuresideandhighvelocityonthelow-pressuresideandvelocitydistribution is not uniform at any radius. Due to this fact, the flow may separate from the 0 suction surface of the blade. Thus, C is less than C and the difference is w2 w2 defined as the slip. Another way of looking at this effect, as given by Stodola, is showninFig.2.4,theimpelleritselfhasanangularvelocityvsothat,relativeto theimpeller,thefluidmusthaveanangularvelocityof2v;theresultofthisbeing acirculatorymotionrelativetothechannelorrelativeeddy.Thenetresultofthe previous discussion is that the fluid is discharged from the impeller at an angle relative to the impeller, which is less than the vane angle as mentioned earlier. Figure 2.3 Pressure distribution on impeller vane. LP ¼ low pressure, HP ¼ high pressure. Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 51 Figure 2.4 Relative eddy in impeller channel. Hence, the slip factors is defined as 0 C w2 s ¼ ð2:3Þ C w2 Forpurelyradialblades,whichareoftenusedincentrifugalcompressors,b will 2 be 908 and the Stodola slip factor becomes p s ¼ 12 ð2:4Þ n where n is the number of vanes. The Stanitz slip factor is given by 0:63p s ¼ 12 ð2:5Þ n When applying a slip factor, the Euler pump equation becomes W ¼sU C 2U C ð2:6Þ 2 w2 1 w1 m Typically,theslipfactorliesintheregionof0.9,whiletheslipoccursevenifthe fluid is ideal. 2.4 PUMPLOSSES The following are the various losses occurring during the operation of a centrifugal pump. 1. Eddy losses at entrance and exit of impeller, friction losses in the impeller, frictional and eddy losses in the diffuser, if provided. 2. Lossesinthesuctionanddeliverypipe.Theabovelossesareknownas hydraulic losses. 3. Mechanical losses are losses due to friction of the main bearings, and stuffing boxes. Thus, the energy supplied by the prime mover to Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved52 Chapter2 impeller is equal to the energy produced by impeller plus mechanical losses. A number of efficiencies are associated with these losses. Letrbethe densityofliquid; Q,flowrate; H,totalheaddeveloped bythe pump; P , shaft power input; H, total head across the impeller; and h, head loss s i i in the impeller. Then, the overall efficiency h is given by: o Fluid power developed by pump rgQH h ¼ ¼ ð2:7Þ o Shaft power input P s Casing efficiencyh is given by: c h ¼ Fluid power at casing outlet/fluid power at casing inlet c ¼ Fluid power at casing outlet/ðfluid power developed by impeller2leakage lossÞ ð2:8Þ ¼rgQH/rgQH ¼ H/H i i Impeller efficiency h is given by: i h ¼ Fluid power at impeller exit/fluid i power supplied to impeller ¼ Fluid power at impeller exit/ðfluid power developed by impeller þ impeller lossÞ ð2:9Þ ¼rgQH / rgQðÞ H þh ¼ H /ðH þhÞ i i i i i i i i Volumetric efficiency h is given by: v h ¼ Flow rate through pump/flow rate through impeller v ð2:10Þ ¼ Q/ðQþqÞ Mechanical efficiency h is given by: m h ¼ Fluid power supplied to the impeller/power m input to the shaft ¼rgQ ðh þH Þ/P ð2:11Þ i i i s Therefore, h ¼h hh h ð2:12Þ o c i v m Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 53 A hydraulic efficiency may be defined as Actual head developed by pump h ¼ h Theoretical head developed by impeller H ¼ ð2:13Þ ðH þhÞ i i The head H is also known as manometric head. 2.5 THEEFFECTOFIMPELLERBLADESHAPE ONPERFORMANCE The various bladeshapes utilizedin impellers ofcentrifugal pumps/compressors are shown in Fig. 2.5. The blade shapes can be classified as: 1. Backward-curved blades (b , 908) 2 2. Radial blades (b ¼ 908) 2 3. Forward-curved blades (b . 908) 2 As shown in Fig. 2.5, for backward-curved vanes, the value of C (whirl w2 component at outlet) is much reduced, and thus, such rotors have a low energy transfer for a given impeller tip speed, while forward-curved vanes have a high valueofenergytransfer.Therefore,itisdesirabletodesignforhighvaluesofb 2 (over908),butthevelocitydiagramsshowthatthisalsoleadstoaveryhighvalue ofC .Highkineticenergyisseldomrequired,anditsreductiontostaticpressure 2 by diffusion in a fixed casing is difficult to perform in a reasonable sized casing. However,radialvanes(b ¼ 908)havesomeparticularadvantagesforveryhigh- 2 speed compressors where the highest possible pressure is required. Radial vanes are relatively easy to manufacture and introduce no complex bending stresses (Fig. 2.6). Figure 2.5 Centrifugal pump outlet velocity triangles for varying blade outlet angle. Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved54 Chapter2 Figure 2.6 Characteristics for varying outlet blade angle. 2.6 VOLUTEORSCROLLCOLLECTOR A volute or scroll collector consists of a circular passage of increasing cross- sectional area (Fig. 2.7). The advantage of volute is its simplicity and low cost. The cross-sectional area increases as the increment of discharge increases aroundtheperipheryoftheimpeller,and,ifthevelocityisconstantinthevolute, Figure 2.7 Volute or scroll collector. Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 55 Figure 2.8 Cross-section of volute casing. thenthestaticpressureislikewiseconstantandtheradialthrustwillbezero.Any deviation in capacity (i.e., flow rate) from the design condition will result in a radial thrust which if allowed to persist could result in shaft bending. Thecross-sectionalshapeofthevoluteisgenerallysimilartothatshownin Fig. 2.8, with the sidewalls diverging from the impeller tip and joined by a semicircular outer wall. The circular section is used to reduce the losses due to friction and impact when the fluid hits the casing walls on exiting from the impeller. 2.7 VANELESSDIFFUSER For the diffusion process, the vaneless diffuserisreasonably efficient and isbest suited for a wide range of operations. It consists simply of an annular passage without vanes surrounding the impeller. A vaneless diffuser passage is shown in Fig. 2.9. The size of the diffuser can be determined by using the continuity equation. The mass flow rate in any radius is given by m ¼rAC ¼ 2prbrC ð2:14Þ r r where b is the width of the diffuser passage, r b r C 2 2 2 r2 C ¼ ð2:15Þ r rbr where subscripted variables represent conditions at the impeller outlet and the unsubscripted variables represent conditions at any radius r in the vaneless diffuser. Assuming the flow is frictionless in the diffuser, angular momentum Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved56 Chapter2 Figure 2.9 Vaneless diffuser. Figure 2.10 Logarithmic spiral flow in vaneless space. Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 57 is constant and C ¼ðÞ C r /r ð2:16Þ w w2 2 But the tangential velocity component (C ) is usually very much larger than the w radial velocity component C and, therefore, the ratio of the inlet to outlet r, C r 2 3 diffuser velocities ¼ . C r 3 2 It means that for a large reduction in the outlet kinetic energy, a diffuser with a large radius is required. For an incompressible flow, rC is constant, and, r therefore, tana ¼ C /C ¼ constant. Thus, the flow maintains a constant w r inclination to radial lines, the flow path traces a logarithmic spiral. AsshowninFig.2.10,foranincrementalradiusdr,thefluidmovesthrough angle du, then rdu ¼ dr tana. Integrating we have u2u ¼ tana logðÞ r/r ð2:17Þ 2 2 Substitutinga ¼ 788 and (r/r ) ¼ 2, the change in angle of the diffuser is equal 2 to 1808. Because of the long flow path with this type of diffuser, friction effects are high and the efficiency is low. 2.8 VANEDDIFFUSER The vaneddiffuserisadvantageouswheresmall sizeisimportant.In thistypeof diffuser,vanesareusedtodiffusetheoutletkinetic energyofthe fluidatamuch higherratethanispossiblebyasimpleincreaseinradius,andhenceitispossible to reduce the length of flow path and diameter. The vane number, the angle of divergence is smaller, and the diffuser becomes more efficient, but greater is the friction. The cross section of the diffuser passage should be square to give a maximum hydraulic radius. However, the number of diffuser vanes should have nocommonfactor withthenumberof impeller vanes.The collectoranddiffuser operateattheirmaximumefficiencyatthedesignpointonly.Anydeviationfrom the design discharge will change the outlet velocity triangle and the subsequent flow in the casing. 2.9 CAVITATIONINPUMPS Cavitation is caused by local vaporization of the fluid, when the local static pressureofaliquidfallsbelowthevaporpressureoftheliquid.Smallbubblesor cavities filled with vapor are formed, which suddenly collapse on moving forward with the flow into regions of high pressure. These bubbles collapsewith tremendous force, giving rise to pressure as high as 3500atm. In a centrifugal pump,theselow-pressurezonesaregenerallyattheimpellerinlet,wherethefluid islocallyacceleratedoverthevanesurfaces.Inturbines,cavitationismostlikely Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved58 Chapter2 tooccuratthedownstreamoutletendofabladeonthelow-pressureleadingface. When cavitation occurs, it causes the following undesirable effects: 1. Local pitting of the impeller and erosion of the metal surface. 2. Serious damage can occur from this prolonged cavitation erosion. 3. Vibration of machine and noise is also generated in the form of sharp cracking sounds when cavitation takes place. 4. A drop in efficiency due to vapor formation, which reduces the effective flow areas. The avoidance of cavitation in conventionally designed machines can be regarded as one of the essential tasks of both pump and turbine designers. This cavitationimposeslimitationsontherateofdischargeandspeedofrotationofthe pump. A cavitation parameter is defined as s ¼ pump total inlet head above c vapor pressure/head developed by the pump or at the inlet flange  2 p V p 1 v 1 s ¼ þ 2 H ð2:18Þ c = rg 2g rg ThenumeratorofEq.(2.18)isasuctionheadandiscalledthenetpositivesuction Figure 2.11 Cavitation limits for radial flow pumps. Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 59 head (NPSH) of the pump. It is a measure of the energy available on the suction side of the pump, and H is the manometric head. The cavitation parameter is a function of specific speed, efficiency of the pump, and number of vanes. Figure2.11showstherelationshipbetweens andN .Itmaybenecessaryinthe c s selectionofpumpsthatthevalueofs doesnotfallbelowthegivenvaluebythe c plots in Fig. 2.11 for any condition of operation. 2.10 SUCTIONSPECIFICSPEED The efficiency of the pump is a function of flow coefficient and suction specific speed, which is defined as 23/4 1/2 N ¼ NQ gðÞ NPSH suc Thus,  h ¼fQ;N suc The cavitation parameter may also be determined by the following equation 3/4 3/4 N /N ¼ðNPSHÞ /H s suc 3/4 ð2:19Þ ¼s c 2.11 AXIALFLOWPUMP In an axial flow pump, pressure is developed by flow of liquid over blades of airfoil section. It consists of a propeller-type of impeller running in a casing. The advantage of an axial flow pump is its compact construction as well as its ability to run at extremely high speeds. The flow area is the same at inlet and outlet and the minimum head for this type of pump is the order of 20m. 2.12 PUMPINGSYSTEMDESIGN Proper pumping system design is the most important single element in minimizing the life cycle cost. All pumping systems are comprised of a pump, a driver, pipe installation, and operating controls. Proper design considers theinteractionbetweenthepumpandtherestofthesystemandthecalculationof the operating duty point(s) (Fig. 2.12). The characteristics of the piping system must be calculated in order to determine required pump performance. This applies to both simple systems as well as to more complex (branched) systems. Bothprocurementcostsandtheoperationalcostsmakeupthetotalcostofan installationduringitslifetime.Anumberofinstallationandoperationalcostsare directly dependent on the piping diameter and the components in the piping system. Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved60 Chapter2 Figure2.12 The duty point of the pump is determined by the intersection of the system curve and the pump curve as shown above. A considerable amount of the pressure losses in the system are caused by valves, in particular, control valves in throttle-regulated installations. In systems with several pumps, the pump workload is divided between the pumps, which together, and in conjunction with the piping system, deliver the required flow. The piping diameter is selected based on the following factors: . Economy of the whole installation (pumps and system) . Required lowest flow velocity for the application (e.g., avoid sedimentation) . Required minimum internal diameter for the application (e.g., solid handling) . Maximum flow velocity to minimize erosion in piping and fittings . Plant standard pipe diameters. Decreasing the pipeline diameter has the following effects: . Pipingandcomponentprocurementandinstallationcostswilldecrease. . Pump installation procurement costs will increase as a result of increased flow losses with consequent requirements for higher head pumps and larger motors. Costs for electrical supply systems will therefore increase. . Operating costs will increase as a result of higher energy usage due to increased friction losses. Some costs increase with increasing pipeline size and some decrease. Because of this, an optimum pipeline size may be found, based on minimizing costsoverthelifeofthesystem.Apumpapplicationmightneedtocoverseveral Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 61 duty points, of which the largest flow and/or head will determine the rated duty forthepump.Thepumpusermustcarefullyconsiderthedurationofoperationat the individual duty points to properly select the number of pumps in the installation and to select output control. 2.12.1 MethodsforAnalyzingExisting PumpingSystems Thefollowingstepsprovideanoverallguidelinetoimproveanexistingpumping system. . Assemble a complete document inventory of the items in the pumping system. . Determine the flow rates required for each load in the system. . Balance the system to meet the required flow rates of each load. . Minimize system losses needed to balance the flow rates. . Affect changes to the pump to minimize excessive pump head in the balanced system. . Identify pumps with high maintenance cost. Oneoftwomethodscanbeusedtoanalyzeexistingpumpingsystems.One consists of observing the operation of the actual piping system, and the second consists of performing detailed calculations using fluid analysis techniques. The first method relies on observation of the operating piping system (pressures, differential pressures,and flow rates), the second dealswith creating anaccurate mathematical model of the piping system and then calculating the pressure and flow rates with the model. The following is a checklist of some useful means to reduce the life cycle cost of a pumping system. . Consider all relevant costs to determine the life cycle cost. . Procure pumps and systems using life cycle cost considerations. . Optimize total cost by considering operational costs and procurement costs. . Consider the duration of the individual pump duty points. . Match the equipment to the system needs for maximum benefit. . Match the pump type to the intended duty. . Do not oversize the pump. . Match the driver type to the intended duty. . Specify motors to have high efficiency. . Monitor and sustain the pump and the system to maximize benefit. . Consider the energy wasted using control valves. Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved62 Chapter2 Figure 2.13 Typical pump characteristics. 2.12.2 PumpSystemInteraction Theactualoperatingpointonthepumpsystemcharacteristiccurveisdefinedbyits interactionwith theoperatingcharacteristics ofthe installedsystem(Fig.2.13). The system characteristics will consist of: . The total static head, being the difference in elevation between the upstream and downstream controls (generally represented by reservoir levels), . The energy losses in the system (generally pipe friction), which are normally expressed as a function of velocity head. . The interaction point of these curves represents the actual operating point (asshownlater),definingthe Head suppliedbythepumpand the Discharge of the system. The efficiency of the pump under these conditionswillalsobedefined. Note that the efficiency of the pump at this operating point is the critical parameterinpumpselectionandsuitabilityforaparticularsystem(Figs.2.14 and2.15). 2.13 LIFECYCLEANALYSIS Overtheeconomiclifeofapumpedsupplysystem,anumberofdesignparameter will change. System behavior for all possible operating environments is needed (Fig. 2.16). Parameters that will change include: Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 63 Figure 2.14 Pump–system interaction point and pump efficiency. Figure 2.15 Selection of pump type. Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved64 Chapter2 Figure 2.16 Variations in demand and operating characteristics. . Seasonal variations in demand. . Water demand increases as the system population expands. . Increasing pipe friction as the system ages. For all operating conditions, it is necessary to maintain pump operation close to peak efficiency. This can be achieved using multiple pumps and timed pumping periods. Copyright 2003 by Marcel Dekker, Inc. All Rights ReservedHydraulic Pumps 65 2.14 CHANGINGPUMPSPEED The most common pump–motor combination employed in water supply operations is a close coupled system powered by an electric motor. These units can only operate at a speed related to the frequency of the A.C. supply (50cycles/s or 3000cycles/min), with the number of pairs of poles in the motor (N) reducing the pump speed to 3000/N revolutions per minute. Pumpsdriventhroughbeltdrivesorpoweredbypetrolordieselmotorsare more flexible allowing the pump speed to be adjusted to suit the operational requirements. Analysis of system operation will require the head–discharge- efficiency characteristic for the particular operating speed. Giventhehead–discharge-efficiencycharacteristicsforspeedN(intabular 0 form), the corresponding characteristics for any speed N can be established as follows:  0 N 0 Q ¼ Q flow points ð2:20Þ N  2 0 N 0 H ¼ H head points ð2:21Þ N 0 h ¼h efficiency points ð2:22Þ The data set for the new pump speed can then be matched to the system characteristics. 2.15 MULTIPLEPUMPOPERATION The most common type of pumping station has two or more pumps operating in parallel. This provides flexibility of operation in coping with a range of flow conditions and allows maintenance of individual units while other units are in operation. Occasionally situations will be encountered where pumps are operated in series to boost outputs. This is generally a temporary measure as any failure of one unit will severely affect system operation. Composite characteristics (head–discharge curves) are obtained by combining the individual curves. The composite curve is employed in the same manner(i.e.,intersectionwithsystemcurve) sanindividualcurve (Fig.2.17). Where pumps operate in parallel, the composite curve is obtained by adding the flow rates for a given head. Where pumpsoperate inseries,thecompositeisobtainedbyadding heads for a given flow rate. Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved66 Chapter2 Figure 2.17 Composite pump characteristics. 2.15.1 NetPositiveSuctionHead Simply, NPSH is the minimum suction condition (pressure) required to prevent pump cavitation. Conceptually, NPSH can be imagined as the pressure drop between the pump inlet flange and the point inside the pump where the fluid dynamicaction,asitleavestheimpeller,causesapressurerise.SufficientNPSH allowsforpumpingwithoutliquidvaporizinginthepumpfirst-stageimpellereye as the fluid pressure drops due to pump suction losses (Fig. 2.18). TheNPSHrequiredisreportedinheadoffluid(beingpumped)requiredat thepumpinlet.Assuch,NPSHrequiredhasunitsoflength(height).Usually,the datum line for pump NPSH is the centerline of the inlet. This is satisfactory for small pumps. For larger pumps, the NPSH requirements reported by the manufacturer should be verified for the datum line being discussed. The NPSH Copyright 2003 by Marcel Dekker, Inc. All Rights Reserved