流体 机械 外文翻译 外文文献 英文文献 离心式和往复式压缩机的工作效率特性

原文EfficiencyAndOperatingCharacteristicsOfCentrifugalAndReciprocatingCompressorsByRainerKurz,BernhardWinkelmann,andSaeidiVIokhatabReciprocatingcompressorsandcentrifugalcompressorshavedifferentoperatingcharacteristicsandusedifferenteificiencydefinitions.Thisarticleprovidesguidelinesforanequitablecomparison,resultinginauniversalefficiencydefinitionforbothtypesofmachines.Thecomparisonisbasedontherequirementsinwhichauserisultimatelyinterested.Further,theimpactofactualpipelineoperatingconditionsandtheimpactonefficiencyatdifferentloadlevelsisevaluated.Atfirstglance,calculatingtheefficiencyforanytypeofcompressionseemstobestraightforward:comparingtheworkrequiredofanidealcompressionprocesswiththeworkrequiredofanactualcompressionprocess.Thedifficultyiscorrectlydefiningappropriatesystemboundariesthatincludelossesassociatedwiththecompressionprocess.Unlesstheseboundariesareappropriatelydefined,comparisonsbetweencentrifugalandreciprocatingcompressorsbecomeflawed.Wealsoneedtoacknowledgethattheefficiencydefinitions,evenwhenevaluatedequitably,stilldon'tcompletelyansweroneoftheoperator'smainconcerns:Whatisthedriverpowerrequiredforthecompressionprocess?Toaccomplishthis,mechanicallossesinthecompressionsystemsneedtobediscussed.Trendsinefficiencyshouldalsobeconsideredovertime,suchasoff-designconditionsastheyareimposedbytypicalpipelineoperations,ortheimpactofoperatinghoursandassociateddegradationonthecompressors.Thecompressionequipmentusedforpipelinesinvolveseitherreciprocatingcompressorsorcentrifugalcompressors.Centrifugalcompressorsaredrivenbygasturbines,orbyelectricmotors.Thegasturbinesusedare,ingeneral,two-shaftenginesandtheelectricmotordrivesuseeithervariablespeedmotors,orvariablespeedgearboxes.Reciprocatingcompressorsareeitherlowspeedintegralunits,whichcombinethegasengineandthecompressorinonecrankcasing,orseparable"high-speed"units.Thelatterunitsoperateinthe750-1,200rpmrange(1,800rpmforsmallerunits)andaregenerallydrivenbyelectricmotors,orfour-strokegasengines.EfficiencyTodeterminetheisentropicefficiencyofanycompressionprocessbasedontotalenthalpies(h),totalpressures(p),temperatures(T)andentropies(s)atsuctionanddischargeofthecompressoraremeasured,andtheisentropicefficiencyr\Athenbecomes:_[h(p ,s)-h(p,T)]门= disch suet suct suct (Eq.1)s[h(p,T)-h(p ,T)](Eq.1)dischdisch suctsuctand,withmeasuringthesteadystatemassflowm,theabsorbedshaftpoweris:.p=—[h(p ,T )-h(p,T)](Eq.2)门 dischdisch suctsuct(Eq.2)mconsideringthemechanicalefficiencyr\A.Thetheoretical(isentropic)powerconsumption(whichisthelowestpossiblepowerconsumptionforanadiabaticsystem)followsfrom:(Eq.3)dischsuctP=m[h(p ,s)一h(p,T)](Eq.3)dischsucttheor dischsuct suctsuctTheflowintoandoutofacentrifugalcompressorcanbeconsideredas"steadystate."Heatexchangewiththeenvironmentisusuallynegligible.Systemboundariesfortheefficiencycalculationsareusuallythesuctionanddischargenozzles.Itneedstobeassuredthatthesystemboundariesenvelopeallinternalleakagepaths,inparticularrecirculationpathsfiAombalancepistonordivisionwallleakages.Themechanicalefficiencyr)A.,,describingthefrictionlossesinbearingsandseals,aswellaswindagelosses,istypicallybetween98and99%.Forreciprocatingcompressors,theoreticalgashorsepowerisalsogivenbyEq.3,giventhesuctionanddischargepressureareupstreamofthesuctionpulsationdampenersanddownstreamofthedischargepulsationdampeners.Reciprocatingcompressors,bytheirverynature,requiremanifoldsystemstocontrolpulsationsandprovideisolationfromneighboringunits(bothreciprocatingandcentrifugal),aswellasfrompipelineflowmetersandyardpipingandcanbeextensiveinnature.Thedesignofmanifoldsystemsforeitherslowspeedorhighspeedunitsusesacombinationofvolumes,pipinglengthsandpressuredropelementstocreatepulsation(acoustic)filters.Thesemanifoldsystems(filters)causeapressuredrop,andthusmustbeconsideredinefficiencycalculations.Potentially,additionalpressuredeductionsfromthesuctionpressurewouldhavetomadetoincludetheeffectsofresidualpulsations.Likecentrifugalcompressors,heattransferisusuallyneglected.Forintegralmachines,mechanicalefficiencyisgenerallytakenas95%.Forseparablemachinesa97%mechanicalefficiencyisoftenused.Thesenumbersseemtobesomewhatoptimistic,giventhefactthatanumberofsourcesstatethatreciprocatingenginesincurbetween8-15%mechanicallossesandreciprocatingcompressorsbetween6-12%(Ref1:Kurz,R.,K.Brun,2007).OperatingConditionsForasituationwhereacompressoroperatesinasystemwithpipeofthelengthLuupstreamandapipeofthelengthLddownstream,andfurtherwherethepressureatthebeginningoftheupstreampipepuandtheendofthedownstreampipepeareknownandconstant,wehaveasimplemodelofacompressorstationoperatinginapipelinesystem(Figure1).Figure1:Conceptualmodelofapipelinesegment(Ref.2:Kurz,R.,M.Lubomirsky.2006).Foragiven,constantflowcapacityQstdthepipelinewillthenimposeapressurepsatthesuctionandpdatthedischargesideofthecompressor.Foragivenpipeline,thehead(Hs)-flow(Q)relationshipatthecompressorstationcanbeapproximatedby(Eq.4)whereC3andC4areconstants(foragivenpipelinegeometry)describingthepressureateitherendsofthepipeline,andthefrictionlosses,respectively(Ref2:Kurz,R.,M.Lubomirsky,2006).Amongotherissues,thismeansthatforacompressorstationwithinapipelinesystem,theheadforarequiredflowisprescribedbythepipelinesystem(Figure2).Inparticular,thischaracteristicrequiresthecapabilityforthecompressorstoallowareductioninheadwithreducedflow,andviceversa,inaprescribedfashion.Thepipelinewillthereforenotrequireachangeinflowatconstanthead(orpressureratio).Figure2:StafionHead-FlowrelationshipbasedonEq.4.Intransientsituations(forexampleduringlinepacking),theoperatingconditionsfollowinitiallyaconstantpowerdistribution,i.e.theheadflowrelationshipfollows:八-H(Eq.5)P=m・一s(Eq.5)门sH叩•const1=^"p^•Qandwillasymptoticallyapproachthesteadystaterelationship(Ref3:Ohanian,S.,R.Kurz,2002).Basedontherequirementsabove,thecompressoroutputmustbecontrolledtomatchthesystemdemand.Thissystemdemandischaracterizedbyastrongrelationshipbetweensystemflowandsystemheadorpressureratio.Giventhelargevariationsinoperatingconditionsexperiencedbypipelinecompressors,animportantquestionishowtoadjustthecompressortothevaryingconditions,and,inparticular,howdoesthisinfluencetheefficiency.Centrinagalcompressorstendtohaveratherflatheadvs.flowcharacteristic.Thismeansthatchangesinpressureratiohaveasignificanteffectontheactualflowthroughthemachine(Ref4:Kurz,R.,2004).Foracentrifugalcompressoroperatingataconstantspeed,theheadorpressureratioisreducedwithincreasingflow.ControllingtheflowthroughthecompressorcanbeaccomplishedbyvaryingtheoperatingspeedofthecompressorThisisthepreferredmethodofcontrollingcentrifugalcompressors.Twoshaftgasturbinesandvariablespeedelectricmotorsallowforspeedvariationsoverawiderange(usuallyfrom40-50%to100%ofmaximumspeedormore).Itshouldbenoted,thatthecontrolledvalueisusuallynotspeed,butthespeedisindirectlytheresultofbalancingthepowergeneratedbythepowerturbine(whichiscontrolledbythefuelflowintothegasturbine)andtheabsorbedpowerofthecompressor.

Virtuallyanycentrifugalcompressorinstalledinthepast15yearsinpipelineserviceisdrivenbyavariablespeeddriver,usuallyatwo-shaftgasturbine.Olderinstallationsandinstallationsinotherthanpipelineservicesometimesusesingle-shaftgasturbines(whichallowaspeedvariationfromabout90-100%speed)andconstantspeedelectricmotors.Intheseinstallations,suctionthrottlingorvariableinletguidevanesareusedtoDrovidemeansofcontrol.Figure3:Typicalpipelineoperatingpointsplottedintoatypicalcentrifugalcompressorperformancemap.Theoperatingenvelopeofacentrifugalcompressorislimitedbythemaximumallowablespeed,theminimumflow(surgeflow),andthemaximumflow(chokeorstonewall)(Figure3).Anotherlimitingfactormaybetheavailabledriverpower.Onlytheminimumflowrequiresspecialattention,becauseitisdefinedbyanaerodynamicstabilitylimitofthecompressorCrossingthislimittolowerflowswillcauseaflowreversalinthecompressor,whichcandamagethecompressor.Modemcontrolsystemspreventthissituationbyautomaticallyopeningarecyclevalve.Forthisreason,virtuallyallmoderncompressorinstallationsusearecyclelinewithcontrolvalvethatallowstheincreaseoftheflowthroughthecompressorifitcomesnearthestabilitylimit.Thecontrolsystemsconstantlymonitortheoperatingpointofthecompressorinrelationtoitssurgeline,andautomaticallyopenorclosetherecyclevalveifnecessary.Formostapplications,theoperatingmodewithanopen,orpartiallyopenrecyclevalveisonlyusedforstart-upandshutdown,orforbriefperiodsduringupsetoperatingconditions.AssumingthepipelinecharacteristicderivedinEq.4,thecompressorimpellerswillbeselectedtooperateatornearitsbestefficiencyfortheentirerangeofheadandflowconditionsimposedbythepipeline.Thisispossiblewithaspeed(N)controlledcompressor,becausethebestefficiencypointsofacompressorareconnectedbyarelationshipthatrequiresapproximately(fanlawequation):2=C5=C2=C5=C6(Eq.6)Foroperatingpointsthatmeettheaboverelationship,theabsorbedgaspowerPgis(duetothefactthattheefficiencystaysapproximatelyconstant):(Eq.7)P=C-H-Q=匕-C-Q3=C-C-C-N3g7 5 C2 7 5 6 7(Eq.7)6Asitis,thispower-speedrelationshipallowsthepowerturbinetooperateat,orveryclosetoitsoptimumspeedfortheentirerange.Thetypicaloperatingscenariosinpipelinesthereforeallowthecompressorandthepowerturbinetooperateatitsbestefliciencyformostofthetime.Thegasproducerofthegasturbinewill,however,losesomethermalefficiencywhenoperatedinpartload.Figure3showsatypicalrealworldexample:Pipelineoperatingpointsfordifferentflowrequirementsareplottedintotheperformancemapofthespeedcontrolledcentrifugalcompressorusedinthecompressorstation.Reciprocatingcompressorswillautomaticallycomplywiththesystempressureratiodemands,aslongasnomechanicallimits(rodloadpower)areexceeded.Changesinsystemsuctionordischargepressurewillsimplycausethevalvestoopenearlierorlater.Theheadisloweredautomaticallybecausethevalvesseelowerpipelinepressuresonthedischargesideand/orhigherpipelinepressuresonthesuctionside.Therefore,withoutadditionalmeasures,theflowwouldstayroughlythesame—exceptfortheimpactofchangedvolumetricefficiencywhichwouldincrea.se,thusincreasingtheflowwithreducedpresstireratio.Thecontrolchallengeliesintheadjustmentoftheflowtothesystemdemands.Withoutadditionaladjustments,theflowthroughputofthecompressorchangesverylittlewithchangedpressureratio.Historically,pipelinesinstalledmanysmallcompressorsandadjustedflowratebychangingthenumberofmachinesactivated.Thiscapacityandloadcouldbefine-tunedbyspeedorbyanumberofsmalladjustments(loadsteps)madeinthecylinderclearanceofasingleunit.Ascompressorshavegrown,theburdenforcapacitycontrolhasshiftedtotheindividualcompressors.Loadcontrolisacriticalcomponenttocompressoroperation.Fromapipelineoperationperspective,variationinstationflowisrequiredtomeetpipelinedeliverycommitments,aswellasimplementcompanystrategiesforoptimaloperation(i.e.,linepacking,loadanticipation).Fromaunitperspective,loadcontrolinvolvesreducingunitflow(throughunloadersorspeed)tooperateascloseaspossibletothedesigntorquelimitwithoutoverloadingthecompressorordriverThecriticallimitsonanyloadmapcurvearerodloadlimitsandHP/torquelimitsforanygivenstationsuctionanddischargepressure.Gascontrolgenerallywillestablishtheunitswithinastationthatmustbeoperatedtoachievepipelineflowtargets.Localunitcontrolwillestablishloadsteporspeedrequirementstolimitrodloadsorachievetorquecontrol.Thecommonmethodsofchangingflowratearetochangespeed,changeclearance,orde-activateacylinder-end(holdthesuctionvalveopen).Anothermethodisaninfinite-stepunloader,whichdelayssuctionvalveclosuretoreducevolumetricefficiency.Further,partoftheflowcanberecycledorthesuctionpressurecanbethrottledthusreducingthemassflowwhilekeepingthevolumetricflowintothecompressorapproximatelyconstant.Controlstrategiesforcompressorsshouldallowautomation,andbeadjustedeasilyduringtheoperationofthecompressor.Inparticular,strategiesthatrequiredesignmodificationstothecompres.sor(forexample:re-wheelingofacentrifugalcompressor,changingcylinderbore,oraddingfixedclearancesforareciprocatingcompressor)arenotconsideredhere.Itshouldbenotedthatwithreciprocatingcompressors,akeycontrolrequirementistonotoverloadthedriverortoexceedmechanicallimits.OperationThetypicalsteadystatepipelineoperationwillyieldanefliciencybehaviorasoutlinedinFigure4.ThisfigureistheresultofevaluatingthecompressorefTiciencyalongapipelinesteadystateoperatingcharacteristic.Bothcompressorswouldbesizedtoachievetheirbestefficiencyat100%flow,whileallowingfor10%flowabovethedesignflow.Differentmechanicalefficiencieshavenotbeenconsideredforthiscomparison.Thereciprocatingcompressorefl'iciencyisderivedn-omvalveefficiencymeasurementsinRef5(Noall,M.,W.Couch,2003)withcompressionefficiencyandlossesduetopulsationattenuationdevicesadded.Theefficienciesareachievablewithlowspeedcompressors.Highspeedreciprocatingcompressorsmaybelowerinefficiency.Figure4:CompressorEfficiencyafdifferentflowratesbasedonoperationaiongasteadystatepipelinecharacteristic.Figure4showstheimpactoftheincreasedvalvelossesatlowerpressureratioandlowerflowforreciprocatingmachines,whiletheefficiencyofthecentrifugalcompressorstaysmoreorlessconstant.ConclusionsEfficiencydefinitionsandcomparisonbetweendifferenttypesofcompressorsrequirecloseattentiontothedefinitionoftheboundaryconditionsforwhichtheefficienciesaredefinedaswellastheoperatingscenarioinwhichtheyareemployed.Themechanicalefficiencyplaysanimportantrolewhenefficiencyvaluesareusedtocalculatepowerconsumption.Ifthesedefinitionsarenotconsidered,discussionsofrelativemeritsofdifferentsystemsbecomeinaccurateandmisleading.REFERENCESKurz.R..K.Brun.2007."EfTiciencyDefinitionandLoadManagementforReciprocatingandCentrifugalCompressors,"ASMEPaperGT2OO7-27O81.Kurz.R.,M.Lubomirsky,2006."AsymttietricSolutionforCompressorStationSpareCapacity."ASMt:Paper2006-90069.Ohanian.S..R.Kurz.2002,"SeriesorParallelArrangementinaTwo-UnitCompressorStation."Trans.ASMEJEngforGTandPower.V)l.124.Kurz.R..2004."ThePhysiesofCentrifugalCompressorPerformance."PipelineSimulationInterestGroup.PalmSprings.CA.Noall,M..W.Couch.2003,"PerformanceandEnduranceTestsofSixMainlineCompressorValvesinNaturalGasCompressionService."GasMachineryConference.SaltLakeCity.UT.第三篇：离心式和往复式压缩机的工作效率特性RainerKurz,BernhardWinkelmann,andSaeidMokhatab往复式压缩机和离心式压缩机具有不同的工作特性，而且关于效率的定义也不同。本文提供了一个公平的比较准则，得到了对于两种类型机器普遍适用的效率定义。这个比较基于用户最感兴趣的要求提出的。此外，对于管道的工作环境影响和在不同负载水平的影响给出了评估。乍一看，计算任何类型的压缩效率看似是很简单的：比较理想压缩过程和实际压缩过程的工作效率。难点在于正确定义适当的系统边界，包括与之相关的压缩过程的损失。除非这些边界是恰好定义的，否则离心式和往复式压缩机的比较就变得有缺陷了。我们也需要承认，效率的定义，甚至是在评估公平的情况下，仍不能完全回应操作员的主要关心问题：压缩过程所需的驱动力量是什么？要做到这一点，就需要讨论在压缩过程中的机械损失。随着时间的推移效率趋势也应被考虑，如非设计条件，它们是由专业的流水线规定，或者是受压缩机的工作时间和自身退化的影响。管道使用的压缩设备涉及到往复式和离心式压缩机。离心式压缩机用燃气轮机或者是电动马达来驱动。所用的燃气轮机，总的来说，是两轴发动机，电动马达使用的是变速马达或者变速齿轮箱。往复压缩机是低速整体单位或者是可分的“高速”单位，其中低速整体单位是燃气发动机和压缩机在一个曲柄套管内。后者单位的运行在750-1,200rpm范围内(1,800rpm是更小的单位)并且通常都是由电动马达或者四冲程燃气发动机来驱动。效率要确定任何压缩过程的等熵效率，就要基于测量的压缩机吸入和排出的总焓(h)，总压力(p)，温度(T)和熵(s)，于是等熵效率门s变为：_[h(p,s)—h(p,T)]门= disch suct suct suct (Eq.1)(Eq.2)s[h(p,T)—h(p,T)]dischdisch suctsuct(Eq.1)(Eq.2)并且加上测量的稳态质量流m，吸收轴功率为:.p=—[h(p ,T)-h(p,T)]门 dischdisch suctsuctm考虑机械效率门m。理论(熵)功耗(这是绝热系统可能出现的最低功耗)如下:P=m[h(p,s)—h(p,T)] (E3)theor dischsuct suctsuct (q.。)流入和流出离心式压缩机的流量可以视为“稳态”。环境的热交换通常可以忽略。系统边界的效率计算通常是用吸入和排出的喷嘴。需要确定的是，系统边界要包含所有内部泄露途径，尤其是从平衡活塞式或分裂墙渗漏的循环路径。机械效率叩m，在描述轴承和密封件的摩擦损失以及风阻损失时可以达到98%和99%。

对于往复式压缩机，理论的气体马力也是由Eq.3给出的，鉴于吸力缓冲器上游和排力缓冲器下游的吸气和排气压力脉动。往复压缩机就其性质而言，从临近单位需要多方面的系统来控制脉动和提供隔离（包括往复式和离心式），以及可以自然存在的来自管线的管流量和面积管道。对于任何一个低速或高速单位的歧管系统设计，使用了卷相结合，管道长度和压力降元素来创造脉动（声波）滤波器。这些歧管系统（过滤器）引起压力下降，因此必须在效率计算时考虑到。潜在的，从吸气压力扣除的额外压力不得不包含进残余脉动的影响。就像离心压缩机一样，传热就经常被忽视。对于积分的机器，机械效率一般取为95%。对于可分机机械效率一般使用97%。这些数字似乎有些乐观，一系列数字显示，往复式发动机机械损失在8-15%之间，往复压缩机的在6-12%（参考1往复压缩机招致号码：库尔兹，R.，K.，光布伦，2007）。工作环境在这样的情况下，当压缩机在一个系统中运行时，管道长度Lu上游和Ld下游，以及管道pu上游的初始压力和管道pe下游的终止压力均被视为常量，在管道系统中我们有一个压缩机运行的简单模型（图1）。图1：管道段的概念模型（文献2：库尔兹.R，M.由罗穆斯基，2006年）。对于给定的，标准管线定量流动能力将在吸入阶段强加压力p,在压缩机放电区强加压力匕。对于给定的管线，压缩机站头部（H）流（。）关系可以近似表述为H=CTI1I—-近似表述为H=CTI1I—-厂二

J1-q+q•q2

up2J-1（Eq.4）其中C3和C4是常数（对于一个给定的管道几何）分别描述了管道两边的压力和摩擦损失（文献2：库尔兹.R，M.由罗穆斯基，2006年）。除去其他问题，这意味着对于带管道系统的压缩机站，头部所需流量扬程是由管道系统规定的（图2）。特别地，这一特点对于压缩机需要的能力允许头部减量，按照规定的方式反之亦然。管道因此将不需要改变头部的流量恒定（或压力比）。图2：建立在4式上的机头流量关系。在短暂的情况下（如包装其间），最初的操作条件遵循恒功率分布，如头部流量关系如下：八・HP=m・一s=const （Eq.5）门sH叩-const1=^"p•Q并将渐进地达到稳定的关系（文献3：奥海宁S.，R.库尔兹，2002年）

在上述要求的基础上，必须控制压缩机输出与系统要求匹配。该系统需求的特点是系统流程和系统头部或压力比的强烈关系。管线压缩机提供了在操作条件经验下的大量变化，一个重要问题就是如何使压缩机适应这样变化的条件，具体的说就是如何影响效率。离心压缩机具有相当大的平头部和流程特点。这意味着压力比的改变对机器的实际流程有重大的影响（文献4：库尔兹R.，20004年）。对于一个恒速运行的压缩机，头部或压力比随着流量的增加而减少。控制压缩机内的流程可以实现压缩机不同的运行速度。这是控制离心压缩机最便捷的方法。两轴燃气轮机和变速电机允许大范围的速度变化（通常是最大速度或更多的40%或50%到100%）。应当指出，被控制的值通常不是速度，但速度是间接平衡由涡轮产生的动力（受进入燃气轮机燃油流量控制）和压缩机的吸收功率。事实上，在过去15年安装的任何离心压缩机在管线服务方面是由调速器来驱使的，通常是两轴燃气轮机。年长的设施和服务设施在其他管线服务有时使用单轴燃气轮机（允许速度90%到100%的变化）和恒速电动机。在这些装置中，吸节流或可变进气导叶用来提供控制方法。图3：典型的管线运行点绘制成的典型离心压缩机性能图。离心压缩机的运行封套受最大允许速度限制，最小流量（涌）和最大流量（窒息或石墙）（图3）。另一个限制因素可能是可用的驱动电源。只有最小流量需要特别注意，因为它被定义为压缩机的一种气