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【讨论】 最近看龙背上的骑兵3攻略,突然想起前一款游戏&&尼尔&&
整理时间: 06:40 来源: 作者:编辑 点击:次
【楼主】 00:11&
最近看龙背上的骑兵3攻略,突然想起前一款游戏&&尼尔&&
&&&&当初因为看不明白英文玩到神秘森林就没玩了,现在想起来去补了剧情,真特么黑啊....妖主角就没一个有好下场
&&&&然后我在A9看到这篇原来人类也没没好下场
&&&&/thread--1.html
&&&&真是踊蛹俚囊缓诘降,据说这算是龙背1还是龙背2某个结局的续作.....
&&&&网友评论 00:40
&&&&白丝长腿7分失足女，一听就很有看头
&&&&网友评论 00:48
&&&&Reply to Reply Post by 杭嘉湖 ( 00:40)
&&&&虽然我没玩过,但是你确定那是女的?
&&&&网友评论 00:50
&&&&一个制作组，音乐质量远高过游戏本身
&&&&网友评论 01:24
&&&&龙背上的婊子，ps2画质十分厉害
&&&&网友评论 01:40
&&&&Reply Post by 桐生夕月 ( 01:24):
&&&&龙背上的婊子，ps2画质十分厉害
&&&&这个制作组做的游戏都是PS2画面...卖点就是剧情和音乐,其他没了
&&&&网友评论 01:46
&&&&尼尔和龙背系列一个世界观
&&&&网友评论 02:11
&&&&龙背上的碧池，我记得CG很不错。
&&&&网友评论 02:15
&&&&那主角的从人设服饰到角色设定都戳中了我的g点
&&&&网友评论 09:15
&&&&Reply Post by rms117 ( 01:46):
&&&&尼尔和龙背系列一个世界观
&&&&世界观的话......平行世界也算同一个世界观吗?
&&&&网友评论 17:55
&&&&Reply to Reply Post by cgnbear ( 09:15)
&&&&尼尔之书的世界观是用龙背骑兵1的D结局之后1000年吧
&&&&网友评论 22:35
&&&&Reply Post by 阿 ( 17:55):
&&&&尼尔之书的世界观是用龙背骑兵1的D结局之后1000年吧
&&&&那个结局就是穿越到平行世界吧
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指导我不敢说，反正我说的应该有意义的，
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剧集名称: 高智能方程式赛车 OP名:I&#39ll Come ED名:Winners(日文) 剧集原作:矢立肇 剧集介绍:主角和雷神慢慢成长直到最后结婚生子...也真佩服日本人的耐力和创意.本部主要为描写一种高科技智能结合赛车的方程式赛车动画,但其中人物刻画得非常有性格和真实,而赛车的经验刺激镜头也绝对爽,堪称经典赛车动画
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淘豆网网友近日为您收集整理了关于Lubrication, Tribology & Motorsport：润滑摩擦学&amp；赛车的文档，希望对您的工作和学习有所帮助。以下是文档介绍：R=5cmR=10cmR=20cmFigure3:OilfilmthicknessversuscrankangleforthetoppistonringofaconventionalEuropean2.0litregasolineengine,at7500rpm,assumingfullyfloodedlubricationR=5cmR=10cmR=20cmOFT(Pm)at-270q10.018.767.38OFT(Pm)at-180q5.515.735.46OFT(Pm)at-90q10.088.857.46OFT(Pm)at0q1.171.842.23OFT(Pm)at90q9.158.076.81OFT(Pm)at180q5.095.425.17OFT(Pm)at270q10.228.937.51OFT(Pm)at360q3.554.004.02FrictionLoss(W)109.Table5:Effectofringradiusofcurvatureonoilfilmthicknessandfriction.LubricantisassumedtobeSAE-15W/opring.Forsuchanengine,weassumethatthetopringhasaheightof1mm,andisasymmetricparabolicring.Weassumearingtensionof1000kPa.Figure4showsthepredictedoilfilmthicknessversuscrankanglefordifferentringradiiofcurvature.(Forthesesimulationsweassumedanenginespeedof18000rpm,aborediameterof94.7mm,astrokeof42.6mmandacon-rodlengthof104mm–thesearethemidvaluesoftheparametersestimatedinTable4.WealsoassumedthatthelubricantwasanSAE-15W/40grade,withabottomdeadcenterlinertemperatureof100qCandatopdeadcenterlinertemperatureof150qC.bustionchamberpressureofaround110bars–theBMEPoftheconventionalenginewas4bars,bustionchamberpressureofjustover30bars,andWrightreportsthattheBMEPofamodernFormula1engineisbetween13and14bar.ordingly).FullyfloodedStarvedFigure5:PredictedoilfilmthicknessunderthetopringofaconventionalEuropeanengine,operatingat7500rpmwithanSAE-15W/40lubricantFinally,etryonpistonassemblyfriction.FortheFormula1engine,usingthesamedataasforFigure4,andassumingatopringradiusofcurvatureof10cm,Table6summarisesthepredictedtopringfrictionpowerloss,andthepredictedminimumoilfilmthickness,fordifferentlubricantsLubricantPowerloss(W)Minimumoilfilmthickness(Pm)SAE-20W/SAE-15W/SAE-10W/SAE-0W/Table6:Effectoflubricantviscositygradeontopringpowerlossandminimumoilfilmthickness(assumingfullyfloodedconditions)Clearly,Table6showsthatreductionsinpistonringfrictioncanbeachievedbymovingtoalowerviscositylubricant.However,thisisattheexpenseoflowerminimumoilfilmthicknesses.Thereisatrade-offbetweenreducedfriction(andgreaterpoweravailabletothewheels)andenginedurability.InTable6,wecanseethatmovingfromanSAE-15W/40lubricanttoanSAE-2500rpm7500rpmFigure6:Con-rodbearingloadsforconventionalEuropean2.0litregasolineengine,at2500rpmand7500rpmFigure7showsthepredictedoilfilmthicknessforbothloadcurvesusingamodifiedversionoftheShortBearingApproximation,whichallowsforlubricantshearthinning30.The“squeeze”effectisalsoincluded.(ThestandardShortBearingApproximationoftenassumesthatitisonlythe“wedge”effectthatisimportantinthelubricationofenginebearings).ThesimulationswerecarriedoutforanSAE-15W/40lubricant.Twoimportantconclusionscanbedrawnfromtheresults.Firstly,forbearingloadsthataredominatedbygaspressure,bustionchamberpressure.Forinertiallydominatedloads,urelsewhere,andtheremaybetwoorthreepositionsaroundthebearingwheretheoilfilmthicknessislow.Inthecalculationscarriedoutabove,thepredictedoilfilmthicknessesforthehigherspeedconditionweregreaterthanthoseforthelowerspeedconditionbecausetheoiltemperaturewasassumedtobethesameinbothsimulations.Inpractice,itislikelythattheoiltemperatureatthehigherspeedconditionislikelytobehigher,sotheminimumoilfilmthicknessislikely,inpractice,tobelowerthanthevaluesshowninFigure7.2500rpm7500rpmFigure7:Predictedcon-rodbearingoilfilmthicknessbasedonloaddatafromFigure6.Inadditiontooilfilmthickness,itisalsoimportanttoknowwhatflowrateofoilisrequiredtolubricatethebearings.Iftheflowrateisnotsufficient,urwhichmayleadtocatastrophicdamagetothebearings.Figure8showstheflowratespredictedbythemodifiedShortBearingApproximation.2500rpm7500rpmFigure8:Flowrate(litres/min)requiredtoavoidlubricantstarvationofcon-rodbearing,basedonloaddataofFigure6Table7summarisestheresultsforminimumoilfilmthickness,frictionloss,andaverageflowrate,fordifferentlubricants(anSAE-20W/50,anSAE-15W/40,anSAE-10W/30andanSAE-0W/20)fortheconventionalEuropean2.0litregasolineengineat2500rpm,andTable8summarisestheresultsat7500rpm.LubricantMinimumOFT(Pm)AverageFrictionPowerLoss(W)AverageFlowRate(litres/min)SAE-20W/50Thesimpleequationsaboveareveryapproximate,andshouldbeusedforidentifyingtrendsonly.mendedforuseinthedesignofbearings.Forthedetaileddesignofbearings,basedonrigidbearingmodels,therearemanyfast,robustmethodsavailable31-34.Forhighpressures(thosegreaterthanabout50MPa)urate.urs,anditisnecessarytosolveboththeReynolds’equationandtheequationthatdescribestheelasticdeformationofthebearingsurfaces.Thisisgenerallyatimeconsumingcalculationthatrequiresanumericaltreatmenteitherbythefiniteelementorfinitedifferencetechniques35-38.ThepaperontheelastohydrodynamiclubricationofjournalbearingsbyFenneretal36isparticularlyilluminating.Theauthorsfindthatunderhighloads,thereisacentralportionofthebearingwheretheoilfilmthicknessisessentiallyconstant,andwherethepressuredistributioniseffectivelythatofaHertziancontact.Inessence,oncethepressureinthebearingexceedsacertainvalue(inthecaseofthebearinganalysedbyFenneretalthispressurewas200MPa),iteffectivelybecameconstant.Thereforethemainconclusionwasthatthepeakpressureinthecon-rodbearingwasmuchlowerthanwouldhavebeenpredictedfromarigidbearingmodel(therigidbearingmodelpredictedapeakpressureof1200MPa,ountpredictedapeakpressureofonly200MPa).Theeffectsofelasticdeformationonfrictionwerenotreported.Concludingthissection,itissuggestedthatalthoughtherearemanysophisticatedmodelsofjournalbearingsinwidespreaduse,ountofthewayinwhichthelubricantviscositymayvaryaroundthebearing.Thevariationoflubricantviscosityaroundthebearing,duetotemperaturevariations,duetothechangingshearrate,andduetothepressurevariationaroundthebearing,areexpectedtohaveasignificanteffectonpredictionsofoilfilmthickness,andfriction,forbothhydrodynamicandelastohydrodynamiclubricationconditions.VALVETRAINTRIBOLOGYLubricationinthevalvetraindiffersfromthatinthebearingsandthepistonassembly,eptedthatthevalvetrainisinthemixedorboundaryregimeformostofitsoperatingconditions.Thisiscertainlytrueasfarasconventionalenginesareconcerned.Measurementsofvalvetrainfrictiontorquefrequentlyshowthat,undernormaloperatingconditions,thevalvetrainfrictiontorqueincreasesastheviscosityoftheoildecreases.Itisalsointhevalvetrainthatfrictionmodifieradditivesarethoughttohavethelargesteffect.Inavalvetrainlubricationmodel,thekinematicsofthecam-followersystemneedtodescribed,andfromknowingthecamandfollowerprofiles,theoilfilmthicknessmaybecalculatedfromelastohydrodynamiclubricationtheory1.Frictionlossesmaythenalsobecalculated.Totalfrictiontorqueinvalvetrainis*TOTAL,whereanapproximateexpressionis:KZCBTOTAL**where*Bisthefrictiontorqueduetoboundaryfriction,andthesecondtermrepresentshydrodynamicfrictioninthevalvetrainsystem(bothfromthecam-followercontactandfromthecamshaftbearings).Inthetermforhydrodynamicfriction,Cisaconstant,Kisthelubricantviscosity(attheappropriatetemperature)andZisthecamshaftangularspeed.Theboundaryfrictiontorquemaybecalculatedbyestimatingtheoilfilmthicknessusingelastohydrodynamictheory,binedsurfaceroughnessofcamandfollower,andassigninganeffectivefrictioncoefficient.essfullybyresearchersfromShell39,Ford40andNissan41,andisfoundtogivegoodagreementwithexperimentalmeasurements.Alowerviscositylubricantwillgiveahigherboundaryfrictiontorque,whereasthehydrodynamicfrictiontorquewillbelower.Frictionmodifieradditivescanbeaddedtothelubricant,andtheseareeffectiveatreducingtheboundaryfrictiontorque.Figure9showsthevalveliftcurveforatypicalEuropean2.0litregasolineengine(withadirectactingbuckettappetvalvetrainsystem).PowerLoss=392.1WIdlingConditionsValveTrainBearingsPistonPowerLoss=1844.9WUrbanConditionsValveTrainBearingsPistonPowerLoss=7461.8WMotorwayConditionsValveTrainBearingsPistonFigure12:EnginefrictionbreakdownforvariousoperatingconditionsforEuropean2.0litreengineOilgradeValvetrainBearingsPistonassemblyTotalSAE-20W/..5SAE-15W/..9SAE-10W/..1SAE-0W/20*410.282.4Table9:ponentfrictiontolubricantviscositygrade,forEuropean2.0litreengineunderurbanoperatingconditions(*NotethattheSAE-0W/20oilisassumedtobefrictionmodified)InTable9,itcanbeseenthatreducingthelubricantviscosityiseffectiveatreducingfrictioninthebearingsandpistonassembly.However,reducingthelubricantviscositycausesanincreaseinthevalvetrainfriction.ThereasonwhytheSAE-0W/paredtotheotheroilsisthatitisassumedthatitcontainsafrictionmodifier(ifitdidn’tcontainafrictionmodifierthevalvetrainfrictionpowerlosswouldhavebeen641.2W).Unfortunately,wedonothavesufficientdetailsofaFormula1valvetrainssystemtoenableustocalculatethetotalenginefrictionofaFormula1engine.However,weknowthatatthemaximumenginespeed,18000rpm,thevalvetrainfrictionislikelytomakeasmallercontributionthanatlowerspeeds.Weestimatedpistonassemblylossesofaround30kW,andbearingfrictionlossesof25kW.At18000rpm,wewouldexpectvalvetrainfrictionlossestobelessthan10kW(andthatthiswouldmainlybehydrodynamicfrictionlossesinthevalvetrain).Figure13summarisesthisveryroughanalysisforaFormula1engineat18000rpm.TotalFriction=65kWValveTrainBearingsPistonAssemblyFigure13:ApproximateestimateofenginefrictioninaFormula1engineat18000rpm(valvetrainfrictionestimateisuncertainduetolackofinformation)ThestrategyforminimizingenginefrictioninaFormula1engineisfairlysimple.Iftheenginehasalotofboundaryfriction,youwoulduseahigherviscositylubricant,withafrictionmodifier,whereasiftheenginehaslessboundaryfriction,youwouldusealowerviscositylubricant(againwithafrictionmodifier).Bylookingatenginefrictionresultsfortheconventionalengine,itispossibletodecreaseenginefrictionby10or20%byusingthecorrectlubricant.Thereareanumberofpapers41-47whichthereadercanrefertoformoreinformationontotalenginefrictioncalculations.Tooptimizepoweroutputfromahighperformanceengine,itisnecessarytochoosealubricantwhichgivesthelowestpossiblefriction.Thisentailschoosingthelubricantviscositywhichgivesthelowestfrictionovertherangeofoperatingconditionsappropriatefortheengine,andchoosinganoptimumfrictionmodifierforreducingfrictioninboundarylubricatedcontacts.CONCLUSIONInthispaperIhaveattemptedtodemonstratethatthephysicalpropertiesofalubricantareextremelyimportantindeterminingboththeminimumoilfilmthicknessinkeylubricatedcontacts,andtheassociatedfrictionlosses.Itisimportanttousemodelsforthepistonassembly,theenginebearings,andthevalvetrain,urateinformationabouthowthelubricantviscosityvarieswithtemperature,shearrate,andpressure.paredaconventionalgasolineenginewithahighperformanceFormula1engine,andshownthatmatchingthelubricanttotheenginecanleadtosignificantreductionsinenginefriction,withconsequentimprovementsintheavailablepowertothewheels.ACKNOWLEDGMENTSTheauthorwouldliketothankSimonDunning,GlynRoperandMarkWakemforusefuldiscussions.TheauthorwouldalsoliketothankShellGlobalSolutions(UK)forpermissiontopublishthiswork.播放器加载中，请稍候...
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R=5cmR=10cmR=20cmFigure3:OilfilmthicknessversuscrankangleforthetoppistonringofaconventionalEuropean2.0litregasolineengine,at7500rpm,assumingfullyfloodedlubricationR=5cmR=10cmR=20cmOFT(Pm)at-270q10.018.767.38OFT(Pm)at-180q5.515.735.46OFT(Pm)at-90q10.088.857.46OFT(Pm)at0q1.171.842.23OFT(Pm)at90q9.158....
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