positive displacement rece pump fundamentals

8/12/2019 Positive Displacement Rece Pump Fundamentals

1/14


2/14

RECIPROCATING PUMPSDEFINITION AND OVERVIEW

By definition, a reciprocating pump is any machine usingreciprocatingmotion to causefluid to bemoved fromonelocationto another. The most common form of reciprocating pump is thepositive displacement type.

This type of pump traps a fixed volume (Figure 1) of fluid anddisplaces it from suction conditions to discharge conditions bymeans of check valves placed in series, at leastone onthe suctionside and at least one on the discharge side. These check valvesensure fluid movement is in one direction from pump suctiontoward the pump discharge. Since a fixed volume of fluid isdisplaced the rate of flow is directly proportional to speed.Capacity can also be increased by using a pump with multipleplungers or pistons. Pump speed and number of plungers/pistonsavailable is limited by mechanical considerations.

Figure 1. Fi xed Di splaced Volume.

As a result of the fixed displaced volume per pump revolutionand the fact that pressure is independent of pump speed andflow rate, there is no capacity-head curve for these machines aswith centrifugal pumps. If a capacity curve were to be drawn, itwould simply be a straight line from zero capacity and speed to

maximumcapacity andspeedlikeshownin Figure2at 100percent(theoretical) volumetric efficiency and another line at actualapplication volumetric efficiency.

Figure 2. Fixed Displaced Volume Capacity Curve.

Within thesphereof reciprocatingpumpstherearea number offundamental configuration differences that warrant explanation.

These include the difference between power and direct actingmachines, the differencebetweensingleand doubleacting pumps,and the differences between plungers and pistons. This basicconfiguration knowledgeis the foundation of the subject.

American Petroleum Institute standard API-674 (1995) andHydraulic Institutestandardscover manyaspectsof pumplimitationsand the features outlined in this tutorial and therefore should beviewed against thebackgroundof thesetwopublications.

Power or Direct Acting Machines

The differencebetween apower pump and a direct acting pumpcan beillustrated by the following statements:

A power pump drives a pumping element(s) through thereciprocatingmotionusing either acrankandslider mechanismorcamshaft to convert rotary motion and power from an electricmotor, engine, or turbine into reciprocating motion and workenergy insidethe fluid endof the pump.

A direct acting pump drives pumping element(s) through thereciprocating motion by use of an integral reciprocating engine(similar toasteamengine) todirectly transfer enginereciprocatingmotion and power into work energy at the fluid end pumpingelement(s). Because of theintegral reciprocating engine, this typepump can be driven by steam, air, or gas.

Bothpower anddirect acting designs can beeither horizontalor vertical in design(Figure3) and thedirectionof reciprocatingmotion is used to define their configuration. Horizontal pumpsoperatewiththereciprocatingmotionin ahorizontal planeparallelto the ground while vertical pumpsoperatewith the reciprocatingmotion in a vertical planeperpendicular to the ground.

Figure 3. Different Pump Configurations.

Di fferences Between Single and Double Acting Pumps

When the pump starts on suction stroke, thepumping elementbegins to withdraw from the pumping chamber (area betweensuction and discharge valves), and pumping chamber pressuredrops below suction manifold pressure. The suction valve thenopensallowingflowtoenter thepumpingchamber, andthenclosesat the end of the suction stroke. The pumping element begins to

re-enter thepumpingchamber onitsdischargestroke, compressingtrapped fluid until pressure in the pumping chamber is abovethatin the discharge manifold causing the discharge valve to open,allowing a fixed volume of fluid to flow into the dischargemanifold. Each time this happens, the pump element is actinguponthe fluid causing pumping. Therefore, pumpsdesigned to dothis onceper complete cycle or revolution aresingleacting, whilethose designs causing this to happen twice per complete cycle orrevolution aredouble acting.

A single actingpump displaces fluid only in onedirection of itsstroke as illustrated in Figure 4. A double acting pump displacesfluid in both directions of its strokeshownby Figure5.

PROCEEDINGSOF THE TWENTY-FOURTH INTERNATIONAL PUMP USERS SYMPOSIUM 200846


3/14

Figure 4. Recipr ocati ng Pump SequenceSingl e Acting.

Figure 5. Reciprocati ng Pump SequenceDouble Acting.

Differences Between Plungers and Pistons

Theterms plunger and piston areoften used interchangeably, asif they were the same thing. This can lead to confusion. It is

importanttounderstandthedifferencebetweenthetwoandusethecorrect term.A plunger type pumping element, Figure 6, is a smooth rod

attachedto the slider mechanism/crosshead, which imparts energyintothepumped fluid to causepressureto bedeveloped within thefluid cylinder. A stationary seal is used around a plunger and theplunger is allowed to move through the seal rings. Plungers areclassifiedas single acting, therefore, when used onadouble actingpump two are required. Plungers are normally used when smallercapacity andhighpressurearerequired.

Figure 6. Plunger.

Pistontype pumping elements, Figure7, consistof acylindricaldiscequipped with aseal at theouter diameter (O.D.) attached toasmooth rod (piston rod). The piston imparts energy into thepumped fluid to cause pressure to be developed within thefluid cylinder. The small smooth rod (piston rod) acts to transferreciprocatingmotionandpower into work energy at the fluid end.

Theseal at thepiston O.D. moves with thepiston. This technologyis conventionally applied to low pressures andhigher flows.

Figure 7. Piston.

COMPONENT OPTIONS

Correctcomponentselection is animportantpart of ensuringlongterm reliability. The hydraulic requirements andthe constituents ofthepump products dictate a number of important design features.Understanding thesefeaturesand selecting theappropriate design

is essential to achievinglong mean time between failure(MTBF).While the pump manufacturer is crucial in advising the correctconfiguration for service, it is prudent to understand the featuresand benefits of some key components. The following provides ageneral overviewof key fluid endcomponents andwherethey arebestsuitedfor use.

Valve Designs

Valve units (suction and discharge) in essence are check typevalves that are opened by differential pressure of the fluid, andallow flow in onedirection only. Themajority of valvedesigns arespring-loaded to rapidly return the valve back onto the seatingsurfaceandminimizethevalveslip (reverseflow throughthevalvebeforeit seats). Themost widely used valve designs areillustratedin Figures 8 through14.

The standardwing guided valve, Figure 8, has wings extendingfromthevalveheadthat guideit asit opensandcloses.Theseatingsurfaces between the valve and valve seat conventionally arebeveledandlapped together to createa tight seal.

Figure 8. Wing Guided Valve Standard.

POSITIVE DISPLACEMENT RECIPROCATING PUMP FUNDAMENTALSPOWERAND DIRECT ACTINGTYPES

47


4/14

This valve is primarily used for higher pressureapplications,up to 10,000 psig (690 barg); it is best suited for clean fluids.However, it can tolerate some suspended solids in the pumpedfluid. Particulates up to 45 micron and in concentrations not toexceed1percentby volumecanbeaccommodated.This valveiswell suited for application pump speeds up to API-674 (1995)andHydraulic Institute speed limitations.

As with the standard wingguided valve, the design in Figure9, also has wings to guideit as it opens and closes; the seatingsurfaces between the valve and valve seat are again beveled.

However, this design has anelastomer/polymer insert imbeddedintothevalveheadthat conforms aroundtrappedsolidstocreateatight seal when thevalvereturns tothevalveseat.Thevalveissuitable for fluids containing particulates up to 45 micron andin concentrations not to exceed 3 percent by volume with a20 to 25 percent pump speed reduction of API-674 (1995)

Table 1 limits.

Figure 9. Wing Gui ded Valve Abrasion Resistant.

The disc valve design, Figure 10, consists of a metal disc

guided by a stem(bolt or studwith wear sleeve) at thecenter asit opens andcloses. Therearetwo seatingsurfaces onthe valveand valve seat, one at the discs outer edgeand one toward thecenter near theguidestem. Thesesurfaces createa tight seal.

Figure 10. D isc Type Valve Standard Center Guided.

This valve type is primarily used for low to moderate pressureapplications, upto4000psig (276barg) andis bestsuitedfor cleanfluids. Thedesigncan toleratesome suspended solidsin the pumpproduct, up to 25 micron and in concentrations not to exceed 1

percent by volume. Well suitedfor application pump speedsuptoAPI-674 (1995) andHydraulic Institute speed limitations.

The design in Figure 11 uses an elastomer/polymer discguided by a stem(bolt or studwithwear sleeve) at thecenter asit opens and closes. There are two seating surfaces on both thevalve andvalveseat, oneat thediscsouter edgeandonetowardthe center near the guide stem, which creates a tight seal.

Figure 11. D isc Type Valve Center Guided Abrasion Resistant.

Thisdesignis primarilyusedforlow to moderatepressureappli-cations, up to 4000 psig (276 barg) and is best suited for fluidscontaining solids up to 45 micron and in concentrations not toexceed 3 percent by volume with a 20 to 25 percent pump speedreduction of API-674 (1995) limits.

The double ported disc valve design, Figure 12, is also a metaldisc guided by a stem/cage at the center as it opens and closes.

Therearetwoseatingsurfacesonboththevalveandvalveseat, oneat the discs outer edgeand one toward the center near the guidestemto createa tight seal.

Figure 12. D isc Type Valve D ouble Port ed.

Theuniquefeatureof this design is that it allows flow aroundthe outside diameter of the valve, as with all the other singleportedvalvedesignspreviously presented, andalso permitsflowthrough the center of the valve near the guide stem. Hence thename double ported. This additional flow area at the discscenter permits higher flow rates for the same valve size.

Primarily used for low to moderatepressure applications, up to3000psig(207barg), it is bestsuitedforcleanfluids. However,thisdesigncan toleratesomesuspended solidsin thepump product up



5/14

to 25 micron and in concentrations not to exceed 1 percent byvolume. Suitable for application pump speeds up to API-674(1995) andHydraulic Institute speed limitations.

As thenameindicates this valveis a ball, Figure13, which isoften guided by an exterior cage as it opens and closes. Theseating surface of the valve is spherical while the valve seat isnormally beveled.This forms aradial linecontactbetween thevalve and valve seat to createa tight seal.

Figure 13. Ball Type Valve.

Largely used in small chemical injectionmetering pumps, thedesign is also well suited for slurry and sludge services. Thisdesigncantoleratesolidsupto250micronin concentrationsnotto exceed 40 percent by volume. Useof this design does haveanegative aspect; dependant upon stroke length and pumpproduct, pump speed should not exceed 50 to 55 percent ofAPI-674 (1995) limits becauseof the weight of the ball.

This elastomeric valve type, Figure 14, is best suited forslurry services (drilling mud, bauxite, borax, and coal). Thedesign consists of an elastomer fastened to valve head and hasstems extending at both ends to guide it as it opensand closes.

The seating surfaces between the valve and valve seat arebeveled.Theelastomer insert conforms around trapped solidstocreatea tight seal when thevalve returns to the valve seat.

Figure 14. El astomeri c Type Valve.

Design variation usingthis principle is available for pressuresbetween 750 psig (52 barg) and 5000 psig (345 barg). Thisdesigncantoleratesolidsupto250micronin concentrationsnot

to exceed 40percent by volume. This design does not havethesame speed limitation requirement as the ball valvedesign.

Stuffing Box Designs and Applications

Reciprocating pumps use various methods of packing as theirsealing methodology. For packing to bereliable it must always bein contact with a lubricant such as the pumped product or anexternal flush. Becauseof this lubricating effect the packing in areciprocating machine will always leak to some extent. Thisleakageis essential in ensuring longpackinglife.

Thepumpvendors experienceis crucial in determining themostappropriate packingoption for the service. This experienceshouldbe used to ensure both long packing life and low leakage rates.Choosing the correct stuffing box configurationfor the fluid beingpumped, its operating conditions, and any EnvironmentalProtectionA gency (EPA) or regulatory requirementsareessential toboth emissions control andreliability. The most common stuffingbox designs available, with their advantages or disadvantages, aredescribedbelow.Thisservesasaguidewhenconsideringthedesignoptions for an application. The pump vendor should assist infinalizingthe designconfiguration.

This simpledesign, Figure15, allows flexibility for various sealtypes. Packingmay besquare, V ring, or nonadjustable designs.

Themajority of thelubricant migratesintopumpedfluid. Thiscon-figurationcan be used ona broad rangeof fluids. It is particularly

suitable for thosefluids that exhibit poor lubricity.

Figure 15. Standard Lubricated Stuffing Box.

Figure16illustrates asimplebox designthat allowsflexibilityfor various seal types andperforms morereliably than astandardlubricated stuffing box in the majority of cases. This designallows lubricant to be fed under the last ring of packing. Thisarea is wherethe lubrication is most needed to resist the slidingwear. With this configuration very little lubricant migrates intothe pumped fluid and the design is suitable for the use of lowpressure drip type lubricators. Packingmay besquare, V ring,or nonadjustable designs.

Figure 16. Alternate Lubricated Stuffing Box.


49


6/14

This type of design, Figure 17, has no manual adjustment; it isself-adjusting as thepackingwears.Thedesigngives longpackingand plunger life and minimal pumped fluid leakage. The designputs lubricant under the last ring of packing as with the previousdesign and allows use of low pressure drip type lubricators.Packing may be squareor V ringdesigns.

Figure 17. Spring Loaded Lubricated Stuffing Box.

The configuration in Figure 18 illustrates the improper use of astandard stuffingbox. This is considered a poor applicationas thehighfriction betweenplungers andpacking causesexcessheatandas a consequence themachinehas short packingandplunger life.

Figure 18. Improper Use of Standard Stuffing Box Used to BleedLeakage to Low Pressure Point.

Figure 19 is an improvement over a standard stuffing box used

to bleed-off leakage to low pressure point although secondarypackingis not adjustable tocompensatefor wear. Lessfrictionandlower temperatures are evident when compared to the standardstuffingboxin Figure18usedtobleed-off pumpedfluidleakagetolow pressurepoint.Thedesignhaslonger packingandplunger life.Packingmay besquare, V ring, or nonadjustable designs.

Figure 19. Stuffing Box with Modified Gland Follower to AllowBleed-Off to Low Pressure Point.

The double gland lubricated stuffing box in Figure 20 isstandard for high pressure, critical (toxic and volatile) services.Averagepacking and plunger life are experienced with negligiblepumped fluid leakageto atmosphere.

Figure 20. Two or Double Gland Lubricated Stuffing Box.

Pumpageleakageiscontrolledby bleed-off tolowpressurepoint(suction, flare, etc.) and independent manual adjustment isrequired for primary and secondary packing glands. Correctadjustment requiresaskilledmechanic. Other featuresinclude: fullsize primary andsecondary packing, positivepackinglubrication,and most of the lubricant migrates toward low pressure bleed-off(suction). Packing may be squareor V ringdesigns.

This design, Figure21, is suitablefor cool water and fluids withcomparable lubricity. It is importantthat thetotal packinglengthisless than the plunger stroke length to ensure lubrication to the lastring of thepackingby thepumpage(self-lubricated). Packingmaybe square or V ringdesigns. Typical packing style and material

usage can be found in Table 1.

Figure 21. Standard Nonlubricated Stuffing Box.

Table 1. Typical Packing Style and M aterial Usage.



7/14

Plunger and Piston Pumping Elements

These components displace the fixed volume of fluid andgenerate pressure and work with the stuffing box seal. Dueconsideration should be given to the needs of the fluid beingpumped, operating conditions, customer requirements, and anyEPA or other regulations for a given application. Pump vendorshaveextensive experiencein this area andshould beconsulted forassistancein finalizing the appropriate design.

Plunger Type Power and Direct

Acting Pump Pumping Elements

These components aresmooth rods that move in and out of thefluid cylinder to cause differential pressure and transmit force todevelop thesepressures insidethe fluid end.

Plungerscanvary in basic design, asshownin Figure22, andaremade of a base metal compatible/suitable for the service with ahard facecoating. Most commonly used coatings include chromeplating, tungsten carbide, nickel/chrome/boride, or one of severalgrades of ceramic. This coating provides wear resistance to thestuffingbox packingseal andextends plunger life.

Figure 22. Plunger Designs.

Piston Type Pumping Elements

These elements consist of a cylindrical disc normally equippedwithaseal of sometypeat theO.D. attached toasmall smoothrod(pistonrod). Pistons areavailable in several designsfor applicationdifferential pressures upto 1500psig (97barg), andall rely onthefluid pumped for lubrication to the piston bearing surfaces andseal. Design selection is dependent on the best O.D. seal for theapplication and ease of maintenance. The body and followerdesigns shownin Figures23through26arethemostcommontypesbecause the piston O.D. seal can generally be replaced without

having to removethe entire piston assembly from the pump.

Figure 23. Piston with Hydraulic Packing Seal.

Figure 24. Piston with Metallic Three Ring Seal.

Figure 25. Piston with Pump Cup Seal.

Figure 26. Piston with Bull Ring and Metalli c or Phenoli c Ring Seal.

Figure 23 piston design has a hydraulic packing seal that isnormally an elastomer and woven fabric (cotton, nylon, aramidfiber, etc.) laminated andcut to formrings. The seal is energizedby the elastomer used and/or swelling, and also by pumpingpressures.This designis well suited for various pumped fluidslikelubeoil, grease, andslopsservices, but it is most popularly usedonwater applications likeboiler feed.

The piston design in Figure24 has a metallic three ring seal, selfenergizedby springaction. It consistsof two outer seal rings andaninner expander ring to ensure an O.D. seal is maintained. The most

commonmaterials arecast-ironandni-resist, but afewapplicationsrequire stainless steels. Long seal life is achieved when used on ametallic liner and pumping afluid withgoodlubricationproperties.

The Figure 25 pump cup seal piston design has a selfenergized cup seal that is enhanced by pumping pressures.Normally cupseals arealaminated elastomer/polymer reinforcedwithfabric (cotton, nylon, aramid fiber, etc.), or can bemadeofpolytetrafluoroethylene (PTFE). This design is well suited forvarious pumped fluids like lube oil, grease, and slops services,and water applications likeboiler feed, just a few examples.

The Figure 26 piston design is the least popular due to thethinner seal cross-sectionandthe additional bull ring part to bemaintained or replaced. Seal rings are typically made of hardplastic or metal, and require a piston liner material with suitablewear resistance. This designrelies solely upon pumpingpressures

to createa seal, and can be used onpumped fluids from lube oil,grease, slops services, and water applications. When correctpiston, seal, and liner materials are used for an application, longseal life can beachieved.

SPECIFICATION CRITERIA

Identifying the servicethat is ideal for a reciprocating pump isnot always a straightforward exercise. Multiple variables must beevaluated and assessed. To make the best choice between pumptypesit isimportanttounderstandhow variouspumptypesbehave.Positive displacement reciprocating pumps are generally used forthefollowingreasons:


51


8/14

Hydraulics

Where the hydraulic requirements of a process system make akinetic or rotary pump unsuitable for the service. Typically, lowercapacities (up to about 600gpm(137 m3/hr) andhigher pressures(upto 20,000 psig (1379barg).

Becauseit is well suitedforlower flow rates.Certain combinationsof flow rate and pressure make centrifugal pumps inherently lessefficient. These applications are around 100 gpm and lower, butespecially under 50 gpm. Centrifugals, do well at higher

flowrates.

Positive displacement reciprocating pump net positive suctionheadrequired(NPSHR) variesasafunctionof flow, whichis deter-mined byspeed. Thelower thepositive displacement reciprocatingpump speed, thelower the NPSHR.

A positive displacement reciprocating pump can be consideredas an alternative to a centrifugal pump operating away frombest efficiency point (BEP). The farther away from BEP acentrifugal pumpoperates, themoreviable apositive displacementreciprocating pump solution becomes. This typically happens atlow flow conditions, moderate to high pressure conditions, orwith elevated viscosity.

Capacity and Pr essure C ontrol

Because of the fixed volume of fluid displacement at agiven speed, a more precise capacity can be achieved. A positivedisplacement reciprocating pump has a constant flow regardlessof pressure; pressure and flow are independent of each other. Acentrifugal pump has varyingflow depending onpressureor head.

Usingareciprocatingpumpvariablecapacity canbeachievedbychangingthepump speed.

A positive displacement reciprocating pump can beused if theapplication has variable pressure conditions. A centrifugal pumpwill beforcedupanddowntheperformancecurvethusvaryingtheflow. This can cause process problems. A positive displacementreciprocatingpumpwill givenear constantflowmakingit possibleto match pump flow rate to the process requirements. Using areciprocating pump variable capacity can beachieved by changing

thepump speed.

When a constant flow is required, a positive displacementreciprocating pump is the pump of choice, especially formeteringapplications.

Reciprocating pumps can be considered self-priming with theright suctionconditions.

Efficiency

Pump types are very different when considering mechanicalefficiency. Pressure changes have little effect on the positivedisplacement reciprocating pump because pressure and flow areindependent of each other. Centrifugal pump characteristicperformancecurves haveinterdependency betweenflow and head.

Energy efficiencyAs a positive displacement reciprocatingpump has a high mechanical efficiency, a smaller driver can beused, resultinginlower power/energyconsumptionversus that of akinetic or rotary pump.

Special Fl uid Requirements

Because of their lower operating speeds, reciprocating pumpscan be used oncertain types of fluidscontainingsolids that wouldquickly erodethe internals of ahighspeed kinetic or rotary pump.

Another major difference between pump types is the effectviscosity has on the capacity of the pump. A centrifugal pumps

flow reduces as the viscosity increases. The positivedisplacementreciprocating pump increases flowdueto theincreased volumetricefficiency caused by higher viscosity liquidsfilling the clearancesinsidethepump fluid end.

Pumps have a tendency to shear fluids as speed is increasedand the centrifugal is a high speed pump. Lower speed positivedisplacement reciprocating pumps are better whenhandling shearsensitive fluids. Shear rates in positivedisplacementreciprocatingpumps vary by design, but they are low shear devices. Pumpmanufacturershaveproprietary information for specific shear rates

and applications andcan provide somerecommendations.

Likeall otherpumpdesigns, reciprocatingpositivedisplacementpumpshavetheir ownset of disadvantages, with themost commonbeingthepulsations.Thepulsatingflow andpressurehavespecificsuction and discharge piping systems requirements to ensureproper pump performance and to protect the piping and othersystemcomponents.

SPECIFICATION METHODOLOGY

As with all pump types (i.e., centrifugal, rotary, screw, etc.)havingcorrectandcompleteapplicationinformationis essential toselectingthemost suitable model andsizeof reciprocatingpositivedisplacementpump.Thefollowingapplication information shouldbeprovidedtothepumpvendor toallowthepumpmodel selection

andsize to bedetermined.

What is the fluid to be pumped, i.e., water, gasoline, triethyleneglycol (TEG), ammonia, etc.?This tells the pump vendor thefluid type (Newtonian, thixotropic, colloidal, or dilliant). Manycommonly pumped fluids are Newtonian, which some pumpvendors have solely establish their pump selection and sizingupon. However, other fluid types have special considerations forpump selection and sizing. The pumped product is used to give ageneral indication of the pump fluid end design and materials ofconstruction required.

Is the fluid to be pumped toxic, flammable, or otherwisehazardous to personnel and/or the environment?This gives ageneral indication of the pump fluid end design, materials ofconstruction, and stuffingbox seal design.

Required delivered capacityA reciprocating pump is apositive displacement type, and as such, it does not have a ratedflow; it only has a given flow at a given speed. Therefore, a rangeof flows can be provided simply by varying the pump speed. Thiscan be achieved by an electric motor with a variable frequencydriveonpower pumps, or by throttling the drivemedia at the inletona direct actingpump.

SuctionpressureThis indicates if astandard pump designcanbeused, orif thepumpsizeneedstobemodifiedtohandle elevatedor high suction pressure (equal to or greater than 5 percent of thedischarge pressure). Also, tells the pump vendor what pressurecategory thesuctionsideof the pump must to berated for.

Discharge pressureUsed to determine pump plunger/pistonsize, frame/rod load to achieve best mechanical efficiency, and

what fluid enddesignpressureis required.

Net positive suction head available (NPSHA)Aids indetermining pump plunger/piston size, pump speed, suction valvetype and size, and what other modifications (if possible) may berequired to provide a low enough NPSHR.

Fluid pumping temperatures: normal, minimum, andmaximumProvides indications of what pump fluid enddesignisrequired, materials of construction, and stuffing box seal designneeded. This also indicates if any pump speed reduction may berequired. For pumping temperatures between 350F (177C) and850F (455C) pumpspeedmust bereducedandaspecial fluidend



9/14

design is required to make allowance for thermal growth tomaintain good alignment between the pumping element andstuffingbox seal.

Any pump speed limits imposed by user specificationsThis isneeded to determine the pump plunger/piston size, and actualpump speed to comply for the application.

Fluid specific gravityThis is used in determining the pumpplunger/piston size, pump speed, and what modifications may berequired for alow enough NPSHR.

Fluid compressibility factor or bulk modulus at pumpingtemperaturesThis is used to determine the pump volumetricefficiency and in turn pump speed for a given plunger/piston sizewithin afluid end.

Fluid vapor pressure at pumping temperaturesNeeded todetermine pump plunger/piston size, pump speed, and anymodifications required for low enoughNPSHR.

Fluid viscosity at pumpingtemperaturesAllowsdeterminationof pump plunger/piston size, pump speed. Starting at 300 SSU(~72cP) pump speed needsto bereduced as indicated in API-674(1995)SecondEditionFigure1andANSI/HI 6.1-6.5(1994). Also,helps to determine what or if modifications may be required forlow enough NPSHR.

Does fluid contain any substances or chemicals that will cause

corrosion and/or erosion within the pump selected?This aids indeterminingthepumpfluid endsize, pumpspeed, pumpvalvetypeandsize, stuffingbox seal design(as requiredfor personnel and/orenvironment protection), and other modif ications needed toprovidelow NPSHR.

As shown in theabove, havingall of this importantinformationhelps determine key aspects of the pump selected and used, inaddition to many important pump performance and constructioncharacteristics, including materials of construction, stuffing boxseal type, andfluid enddesign.

RECIPROCATING PUMP EFFICIENCY

While centrifugal pump efficiency is understood in terms ofleakage, mechanical, and hydraulic losses, the efficiency of areciprocatingmachineissomewhat different. Themain constituentsof reciprocating pump efficiency arecovered below.

Volumetric Efficiency

As shown in the following equation reciprocating pumpvolumetric efficiency (VE) can be determined with reasonableaccuracy(within 1percent), if all factorsareknown.AlsoevidentisthatVE is dependant upon the fluid compressibility, application pressure,pump C/D ratio (pumping chamber clearance to displacement ratio),andpump valveslip. Therefore, sincefluid compressibility, pump C/Dratio, and pump valve slip are known by fluid properties and pumpdimensions and characteristics, the actual fixed displaced volume percompletecycle (rpm) is dependantuponpressureandnot pump speed.

where:P =Differential pressure (psig) =PD PS =Compressibility factor of fluid to be pumped at pumping

temperature reciprocal (inverse) of fluid bulk modulusatpumping temperature

=Ratio of total volume between the suction and dischargevalves inside the pumping chamber when the plunger(or piston) is at full forward stroke divided by the plunger(or piston) displacementvolume(areastrokelength); alsocalledC/D ratio

VL =Valve loss or VE loss from fluid slippage back past thepump valves before they can close and seal. This will

vary between 1 to 5 percent based upon pump speed andvalve design. In general, most pump designs willtypically havea 3 percent loss.

Mechanical Efficiency

The mechanical efficiency (ME) of a reciprocating machine,as previously mentioned, is an important consideration. Thissection outlines the influencing factorswith respect to mechanicalefficiency andillustrates bothpower anddirect acting pumps.

Power PumpsAs shown in Figure 27, mechanical efficiency varies as a

functionof frame load, which is simply the pressure applied timesthecross-sectional area of the plunger/piston. Percentageof frameloadis actual frameloadfromtheapplicationdividedby theframeloadrating established by the pump vendor. Higher frame loadingincreases ME. Therefore, the plunger/piston should be sized toprovide as high as possible frame load within the pumps frameload rating to achievethe highestpossible ME.

Figure 27. Mechanical Effi ciency Versus Frame Load.

Power pumpshavemechanical efficienciesupto87percentwithsleeve type power end bearing pumps, and up to 90 percent forroller typepower endbearingpumps. Contributingtothelossesarethe pump bearings, stuffing box seals, and the pump valves.Additional losses fromdrive components (i.e., driver, belts, gears,couplings, etc.) should be determined and accounted for in theoverall efficiency of the completepump unit.

Direct Acting Pumps

Mechanical efficiency or pump efficiency is a ratio of forceapplied to the fluid pumped by the piston/plunger by the forcetransmittedby thedrivepistonfromthedrivemedia(steam, gas, orair), minus losses due to friction fromdrive piston drag, stuffingbox packing, andpumpvalves. This ratio is basedupondifferentialpressures at both the pumps drive and fluid ends, as determinedusing the followingequation.

where:AL =Cross-sectional areaof fluid piston/plungerpL =Differential pressure acrossthe pump fluid endADR =Cross-sectional areaof fluid piston/plungerpDR =Differential pressureacross the pump driveend

PUMP AND SY STEM INTERACTION

Onesignificant factor in thesuccessful operationof any pump isanappropriately designed pipingsystem. A poorly designed systemcan cause problems such as:

Fluid flashingEntrained gases in the fluid coming out whenpressurein pipingor pump falls below fluid vapor pressure

CavitationFree gases in a fluid being forced back into thefluid. These implosions cause severe pressure spikes that pit anddamage pump internal parts.


53


10/14

Piping vibrationCan result from improper piping support,cavitation, or normal reciprocating pump hydraulic pulses

Noisy operationMost presentwhen pump is cavitatingReducedcapacityCanresult fromfluid flashing. If it is, this isan indication that the pumping chambers are filing up with gasesor vapors.

These factors contribute to reduced pump life, and are apotential hazard to personnel and associated equipment. It ispossible to fracture a fluid cylinder and/or pipingand damage thepumpdriveendinternalswith highpressuresurgesoccurringwhenfluid is flashingor cavitating.

The following basic piping guidelines represent a combinationof Hydraulic Instituterecommendations and criteria established byexperienced designersof systems containingreciprocating pumps.While pump manufacturers cannot assume responsibility for thepiping system into which the pump is installed they can providevaluable guidancethat can aid systemdesigns.

Pump Suction Piping System

Piping should be laid out so no high points occur where vaporpocketsmay form. Vapor pocketsreducethe effectiveflow areaofthe pipe and consequently make pump priming and operationdifficult. Vent any unavoidable highpoints andprovidegauge anddrain connectionsadjacent to pump.

Li ne Size

Many pumpproblems result froma suctionlinethat is too smallin diameter, or too long. Suction piping should be as follows toprovide a smooth transition of fluid flow and result in reducedpiping frictionlosses:

Be shortand direct

Be one to two sizes larger than pump suction connection. Useeccentric type pipereducers at pump with flat side up to avoid apossible vapor pocket.

Contain a minimum number of turns. Accomplish necessaryturns withlong radius elbowsor laterals.

Figure28 containsachart toaid in thedesignof suction piping.

From this figure determine the optimum suction velocity forthe appropriate machine and use this information in Figure 29to determine the appropriate pipe size. This pipe sizing chart isbased on Schedule 40 piping and the values should be adjustedaccordingly for heavier schedules.

Figure 28. Maximum Recommended Suction Line Velocity forPump Type.

Figure 29. Flow Velocity in Suction Pipe Standard Weight(Schedule 40) Steel Pipe Selection Curve.

Thesuction linevelocity is based on an acceleration head of 0.7feet per foot of suction line length and the acceleration headequation fromtheHydraulic InstituteStandard. For dischargeline,a velocity not exceeding three times the suction line velocity isconsideredgoodpractice.

Acceleration Head

After selecting the line size from above, it is necessary tocalculate the total acceleration head (Ha) in the suction line.As with centrifugal machines the NPSHA in the system mustalways exceed pump NPSHR plus piping friction losses, but withreciprocating pumps a further consideration must be made withrespect to acceleration head.

Fluid insidethepump is accelerated anddecelerated as aresult ofthe reciprocating motion and suction valves opening and closing.

This accelerated and decelerated pulsation phenomenon is also man-ifested within thesuction pipe. Theenergy requiredtokeep thefluidin thesuction pipefromfalling below vapor pressureis called accel-eration head. The key influencing factors of the pump are pump

speed, plunger size, valvespringloadandspringrate, valvelift, valvepassagearea, cylinder passageconfiguration, suction manifold con-figuration, andpower endconnectingrodcenter to center of bearingdistanceto throw radius ratio. While the key influencing factors forthesuctionpiping areactual suctionpipelength, mean flow velocityin suction line, and fluid being pumped. If sufficient energy is notavailable, problems such as fluid flashing, cavitation, pipingvibration, noisy operation, reduced capacity, and reduced pump lifecanoccur.TocalculatetheHa requiredtoovercomethis phenomenon,use thefollowing empirical equation:

where:Ha =Head in feet (meters) of liquid pumped to produce required

accelerationL =Actual suction pipelength in feet (meters) not equivalent

lengthV =Mean flow velocity in suctionlinein feet per second(m/s)N =Pump speed in rpmC =Pump constant factor of

0.200for simplex double acting0.200for duplex single acting0.400for simplex single acting0.115for duplex double acting0.066 for triplex singleor double acting0.040forquintuplex single ordouble acting



11/14

0.028for septuplex single or double acting0.022for septuplex single or double acting

g =Accelerationof gravity =32.2 feet/sec2 (9.8m/s2)k =Liquid factor of

2.5 for highly compressible hydrocarbons2.0 for mosthydrocarbons1.5for water, amine, glycol1.4 for deaerated water

Since this equation is based on ideal conditions of a relatively

short, nonelastic suction line, calculated values of Hashould beconsideredasapproximationsonly.

Notes:1. As pump speed (N) is increased, mean flow velocity (V)

also increases. Therefore, acceleration head (Ha) varies as thesquareof pump speed.

2. Acceleration head varies directly with actual suction pipelength(L).

3. Acceleration head is a suction piping system factor thatmust be accounted for by the piping system designer. Pumpmanufacturers cannot account for this in their pump designsbecause of the large variety of applications and piping systemspumpsareinstalled in.

4. If accelerationhead is ignored or miscalculated, signif icantpump and piping systems (suction anddischarge) may result.

Net Positive Suction Head

Net positive suction head (NPSH) for positive displacementreciprocating pumps is normally expressed in pressureunits (psi,kPa, Bar) since a significant portion of pump NPSHR is thepressure requirement needed to push the suction valve from itsseat, to overcome friction losses and acceleration head within thefluid cylinder, rather than an energy per unit mass (or head)requirement.

Expressed in pressure units it becomes independent of fluiddensity. NPSHR pressure units are not expressed as gauge(aboveatmospheric pressure) norabsolute(aboveabsolutezero),

but a pressure measurement abovevapor pressure.NPSH is divided into two important aspects: what is available(NPSHA) fromthesuction vessel and piping, and what is requiredby the pump (NPSHR).

NPSHA

Suction system NPSHA is the pressure provided above fluidvapor pressure to ensure the pump is provided with the NPSH itrequires, plus an adequate amount of pressureto overcome pipingfriction losses and acceleration head generated by the pump. Todeterminethe necessary minimum, total NPSHA must beequal toor greater than pump NPSHR +pipe friction losses +accelerationhead. If suction system design cannot achieve this requirement,NPSHA canbeincreasedby oneormoreof thefollowingremedies.

1. Increase suction pipediameter to makea larger volume of fluidavailable adjacent to the pump suction connection and reducesuction piping flow velocity, which will reduce acceleration headandpiping frictionlosses.

2. Shorten suction pipinglength by providing a more direct routebetween the pump andsuctionvessel, relocatethe pump closer tothe suction vessel, or some combination of both. Again, reducingacceleration head andpiping frictionlosses.

3. Install asuctionpulsation dampener or stabilizer adjacentto thepumpsuctionconnection. This will reducethepulsatingfluidmassinertia effect by changing the suction pipinglength portion in the

acceleration head equation to approximately 20 to 25 times thesuctionpipediameterfor bladder types, andtoapproximately eightto 10 times the suction pipe diameter for larger volume flowthroughtypes.

4. I ncrease available static head by elevating the suction vessel,raisingthefluid level in thesuctionvessel, orpossibly loweringthepump location.

5. Cooling thefluidbeingpumpedafter it leaves thesuctionvessel

to lower the vapor pressure, now making this pressure reductionaportion of the NPSHA.

NPSHR

Pump NPSHR for a reciprocating machine is not analogous tothat of a centrifugal machine. Understanding the differencebetween the two types of NPSHR is fundamental in ensuring thatthe pump operates with the correct NPSH margin.

NPSHR for areciprocatingmachineis definedasthepressureatwhich thepumphasexperiencedalossof nomorethana3percentreductionin capacity toobtain satisfactoryvolumetric efficiency ata specific pressure and speed, and relates to pump suction valvelosses, pump suction manifold acceleration head, plus pumpsuction manifold and pumping chamber frictional losses.Conversely for a centrifugal pump NPSHR relates to 3 percent

reductionin head at a constantflow.NPSHR for a reciprocating pump is generally determined bytestingoncool water, anda 3 percentlossin capacity is anindicationthat the pump is starting to or has entered into a cavitational statecausing thereductionin capacity.

Figure30 is atypical NPSHR curvefor a4 inchstrokesingleacting triplex power pump, which has a range of plunger sizesavailable with a given fluid cylinder size. As can be seen, NPSHRvaries as afunction of plunger size and pump speed within agivenfluid cylinder size. The key influencing factors are pump speed,plunger size, valve spring load and spring rate, valve lift, valvepassage area, cylinder passage configuration, suction manifoldconfiguration, and power end connecting rod center to center ofbearing distance to throw radius ratio.

Figure 30. Typical NPSHR Curve.

It has been suggested that some value above the pumpmanufacturers statedNPSHR beusedbysystemdesignerstoavoidthe possibility of operating the pump in a cavitational state.Suggested values for this margin ranges from103 percent to 125percent of the NPSHR, and varies based upon the type of fluidbeingpumped.Thelower thefluid specific gravity, the higher thissafety factor margin becomes.

Design constraints placed on the system designer can beaccommodated by modifying a reciprocating pump design. Thefollowing factors can be adjusted to provide a more suitableNPSHR values.


55


12/14

1. Reduce pump speeda. Increasetheplunger orpistonsizewithinagivenmodel and

strokelength.b. For agiven strokelength, increasethequantity of plungersor

pistons, i.e., use a triplex pump instead of a duplex, or a quintuplexinstead of atriplex, etc.

c. Change to a larger (longer) strokelength pump.

2. Change the type of suction valve used. Lightweight valvesgenerally require less NPSHR due to the reduction in valve

cracking pressure.

3. If reduced pump speed is possible and changing valve type isnot, NPSHR could potentially be reduced by using a light (lessspring force) suction valve spring, or if pump operating is slowenough, no suction valve springat all.

Ultimately,usingabooster pumptoprovideampleNPSH shouldbe considered if the system constraints dictate. Such a pumpshould be installed adjacent to suction supply vessel, have anNPSHR less than total suction system NPSHA, and have adischargeheadatleast20percentgreater thanpositivedisplacementreciprocating pump NPSHR +pipe friction losses +accelerationhead. A suction pulsation dampener or stabilizer adjacent to thepositive displacement reciprocating pump suction connection isalso appropriate to protect the booster pump from the pulsatingfluid massinertiaof thepositive displacementreciprocatingpump.

Pump D ischarge Pipi ng System

Listed below are the fundamental requirements for a dischargepipingsystem.

Piping should not be smaller than pump discharge connection,and should:

Be as short and direct as possible.Be oneto two sizes larger than pump dischargeconnectionwithincreasersused atpump.

Contain aminimumnumber of turns.Accomplishanynecessaryturns withlong radius elbowsor laterals.

Have a fluid velocity not exceeding three times maximum

suction linefluid velocityshownin Figure29 for pump crankshaftoperatingspeed.

Be providedwithgaugeanddrain connectionsadjacenttopump.

All positivedisplacementreciprocatingpumpsdeliver fluid andbuild pressure until action is taken to control and stabilize thepumps work or a failure occurs. To protect pump, piping, andpersonnel from hazards associated with operating a positivedisplacement pump against a dead head a safety relief valveshould always beprovidedbetweenthepump anddischargevalve.

The safety relief valve should be sized to pass the entire pumpcapacity andthecrackingpressureshould beset at 10percentoverthespecifiedworkingdischargepressureandhaveanaccumulationpressurenot exceeding110 percent of cracking pressure.

Thesafety relief valveoutlet connection should ideally bepipedback to the suction supply vessel. Piping back to the suction pipecan causediscontinuitiesin thesuction pipeflow that can result inpoor pump operation and damage. Should it become necessaryto pipe the safety relief valve back to the suction piping, theconnection into the suction pipe should be a minimum of 10suction pipe diameters in length back toward the suction supplyvessel away fromthepumpsuctionconnection. This will allow anyflow discontinuity created by the relief valve flow into the suctionpipeto besmoothed out by time andviscouseffect.

A discharge bypassline frompump discharge piping back to thesuctionsupply vessel permits lubrication toreachcritical pump anddrivepartsduring startupwithoutsubjectingthemtohighloads and

allowsall fluidcylinder pumpingchamberstobecomefully primed.A bypasslinewithashut-off valve should beinstalled in dischargepiping between pump and check valve back to suction supplysource, not intothepump suction lineto preventflow discontinuity.

Install a dischargecheck valve beyond thebypassconnection toprotect pump from discharge system pressure during pump idleperiods andpump startup.

Discharges piping dead ends are to be avoided or providedwith dampening device. This typeof featurecanberesponsible forundesirable pipingharmonicsandcan contributeto elevated levels

of vibration andnoise.For someservicesthenatural pumppressureor flowfluctuationsmay not be appropriate. In these cases it is prudent to use apulsation dampener for theinstallation. For maximumeffectivenessthe dampener should be mounted adjacent to the pump fluidcylinder. Recommendations for dampener size and type can beobtained from dampener manufacturers based on details of pumptype andsize, service conditions, andpipingsystem.

Install flanges or unions as close to the pump as practical toallow for fluid cylinder removal during maintenance.

Shut-off valvesare requiredin both suction anddischarge linesto isolate pump when maintenance is required. They should beoffull openingdesign, suchas agatevalve.

When connecting two or more pumps to a common suctionand/or dischargelineexercisecaretopreventamutually reinforcing

pressure wave from occurring during operation. This can beachieved by adding the capacities of all pumps that will operatesimultaneously to determine line velocities for sizing pipe andcalculatingtheaccelerationhead.Thebestway toavoid amutuallyreinforcing pressure wave is to install independent suction anddischargelinesto each pump.

Figure31gives an example of the recommendationsoutlined intheprevioussectionfor an appropriate pump piping system, whileFigure 32 provides an example an inappropriate pump pipingsystemconfiguration.

Figure 31. Pipi ng System Appropri ate Design.

Figure 32. Piping System Inappropriate Design.



13/14


57

APPENDIX A

Troubleshooting pump/systeminteraction problems can be found inTableA-1.

Table A-1. Troubl eshooti ng Pump/System Interacti on Probl ems.


14/14


CONCLUSION

Reciprocating pumps are widely used in many industries.Correctlyselectingtheequipmentwith thecorrectconfigurationofoptions is essential to providing and operating reliable machines.

The pump and system interaction is also a vital part of reliableoperation. Thepumpcannotbeviewedin isolation; how it interactswith the system can also causepoor operating life andprematurefailures.

REFERENCES

ANSI/HI 6.1-6.5, 1994, Reciprocating Power Pumps forNomenclature, Definitions, Application and Operation,American National Standards Institute, Washington,D.C./Hydraulic Institute; Parsippany, NewJersey.

API Standard 674, 1995, Positive Displacement PumpsReciprocating, SecondEdition,AmericanPetroleumInstitute,Washington, D.C.

BIBLIOGRAPHY

ANSI/HI 6.6, 1994, Reciprocating Pumps Tests, AmericanNational Standards Institute, Washington, D.C./HydraulicInstitute; Parsippany, New Jersey.

ANSI/HI 8.1-8.5, 1994, Direct Acting Steam Pumps for

Nomenclature, Definitions, Application and Operation,American National Standards Institute, Washington,D.C./Hydraulic Institute; Parsippany, NewJersey.

Binder, R. C., 1943, Flui d Mechanics, New York, New York:Prentice-Hall, Inc.

Henshaw, T. L., 1987, Reciprocating Pumps, NewYork, NewYork:VanNostrandReinhold Company.

Karassik, I., Messina, J., Cooper, P., and Heald, C., 2001,PumpHandbook,Third Edition, NewYork, NewYork: McGraw Hill.

Miller, J. E., 1987, The Reciprocating PumpTheory, Design, andUse,NewYork, NewYork: John Wiley & Sons, Inc.

Ordway, E. P., 1921, Union Engineeri ng HandbookPumpingMachinery, Ai r Compressors, and Condensers, First throughEleventh Editions, Battle Creek, Michigan: Union SteamPumpCompany.

ACKNOWLEDGEMENTS

Dedicated to those special people who have shared theirknowledge, time, and friendship throughout the years: Terry L.Henshaw, P.E.; Harold Hal Wilde, P.E.; Reed Nelson, M.E.;Galen C. Peck; Walter J. Noga; and JamesA. Cripe, M.E.

positive displacement rece pump fundamentals

Documents