Cfd Design And Analysis Of A Compact Single-spool Compressor For A .
CFD DESIGN AND ANALYSIS OF A COMPACTSINGLE-SPOOL COMPRESSOR FOR A HEAVYTRANSPORT HELICOPTER'S POWERPLANTA. Startsev*, Yu.Fokin**, Eu. Steshakov**** ** ***Central Institute of Aviation Motors (CIAM), Moscow, Russia *aerora@bk.ruKeywords: thermodynamic cycle, turbo-shaft engine, axial-centrifugal compressor, CFD,,AbstractPaper concerns optimal thermodynamiccycle of 9000 HP turbo-shaft engine for a twinengine power-plant of heavy helicopter. Thecycle feasibility is approved through design ofthe engine core and detailed CFD design andanalysis of its critical element – compressorPaper begins with estimating the corerelevant overall pressure ratio (OPR) as atrade-off between specific power and fuelconsumption taking into account thermal andstrength limitations. More adequately OPR isdefined from considering given shaft power (SP)as a maximum work that would be obtained inexpanding gas from core exit temperature andpressure to ambient pressure in an imaginarypower turbine. Obtained analytical expressionsare reduced to dependence between compressorefficiency ( c) and OPR, so that optimum OPRis 18, if 82.5% compressor isentropic efficiencycould be achieved.The second part of the paper outlinesdetailed CFD design of the compressor. Thisone-spool high-speed compressor consists of atwo-stage low pressure axial compressor (LPC)followed by a rear centrifugal stage – highpressure compressor (HPC).NomenclatureSP, HP (SHP) – shaft power, shaft horse powerCSP – core shaft powerQ – burner heat additionOPR – overall pressure ratioTET – turbine entry temperatureLPC – axial low-pressure compressorHPC – centrifugal high-pressure compressorLPC PR – pressure ratio of LPCHPC PR – pressure ratio of HPCCT PR – pressure ratio of core turbine c isen, c poly – compressor efficienciesT3 – compressor discharge temperature e isen, e poly – core turbine efficiencies1 IntroductionThe modern engine’s trend towards higherspecific power and reduced specific fuelconsumption leads to aggressive designs: fewercompressor and turbine stages, higher pressureratio and minimal axial length of bladed rowsand spacing between them.Certainly, the next generation enginerequires higher isentropic efficiency of itscomponents. Nevertheless, actual design is atrade-off between level of complexity andcomponent isentropic efficiency.NASA Subsonic Rotary Wing projectconsiders for a future 7500 and 12000 HP-classrotorcraft engines the value of cycle OPR up to40 and turbine entry temperature (TET) equal to3000F (1670K) (see [1]). These cycleparameters are high and cause a number oftechnical challengers for compressor and enginedesign: aerodynamics of low corrected flow inaft stages, strength-of-material and coolinglimitations at high compressor dischargetemperatures, two-spool architecture and otherconstruction complexities.1
STARTSEV A., FOKIN Yu., STESHAKOV Eu.A 9000 HP turbo-shaft engine considered inthis paper is more conservative in OPR. Theengine cycle parameters and shaft power aregiven in Table 1 and compared with thoseproposed by MTU for a similar engine (see [2]).Table 1 Comparison of turbo-shaft engine dataOPRTETSPCIAM18:11650 6.6 MWMTU19 – 21:11800 6 – 7.5 MWMass flow-rate equal to 22.3 kg/s has beenchosen to provide 6.6 MW shaft power. Coreengine includes the following components:compressor with minimum number of stagesand variable stator vanes, short annularcombustor and one-stage axial turbine withcooled turbine stator/rotor blades made ofnickel-base alloys.One of the core design targets iscompactness and simplicity of architecturebased on current level of technology (see Fig.1).Fig.1 Core of 9000 HP turbo-shaft engineCore is single-spool with high rotationalspeed. Core turbine is one-stage and supersonic.Compressor is also supersonic with only onevariable stator vane (namely, LPC inlet guidevane). Stage number is minimized throughaxial-centrifugal configuration of compressor.Flow-path of LPC is tailored so thatintermediate S-shaped duct between LPC andHPC is practically absent.Importance of compressor design has beenshown in paper [3]. Paper [3] outlines optimumengine configurations for light and mediumrotorcrafts where compressor flow-rate andLPC/HPC pressure ratios appear as the principaldesign variables of the engines.Valuable overview of different compressorconcepts for a small 40kW – 100kW classturbo-shaft engine has been given in [4]. Sixcompressor concepts have been assessed. Threestage 6.5:1 pressure ratio compressor (two axialstages and diagonal stage) has been chosen asthe best configuration. It is interesting to notethat two-stage axial-centrifugal compressor withthe same pressure ratio has been rejected. Thereason was that “aerodynamical matchingbetween an axial stage and a radial stagerequires a tuning of the blade tip speeds andthus the blade loading of both stages”.Only few patents disclose proportionbetween axial stages pressure ratio (LPC PR)and centrifugal stage pressure ratio (HPC PR) ofan axial-centrifugal compressor. The PWCpatent [5] proposes single stage axial LPC (LPCPR equals 1.66:1, isentropic efficiency equal to0.87) and centrifugal HPC (HPC PR equals6.04:1 with isentropic efficiency equal to0.829). It is interesting to note that OPR of thistwo-stage compressor is 10:1 and isentropicefficiency is equal to 0.82 (rather high value).On the basis of these data, it looksreasonable to consider as optimum 1:3 ratiobetween LPC PR (with moderately loaded highaspect ratio stages) and HPC PR (with highpressure centrifugal impeller). In this caseimpeller discharge absolute flow is supersonicrequesting careful design of radial bladeddiffuser and outlet system. Returning to [4], onecan read that “real aerodynamic challenge is thedesign of the stator of diagonal stage. The statorhas to provide a very high flow turning at a highinlet Mach number. At the same time, the statorsystem must decrease the meridional Machnumber to values around 0.2 in order tominimize the total pressure loss over theburner”. Carefully designed outlet system ofcentrifugal stage proposed in this paper consistsof a double-row bladed diffuser, de-swirl vanesand pre-diffuser of combustion chamber. Machnumber at the outlet of pre-diffuser is 0.143(with a 3.24 flow swirl).2
CFD DESIGN AND ANALYSIS OF A COMPACT SINGLE-SPOOL COMPRESSOR2 Choice of thermodynamic cycleparametersCore engine configuration includes thefollowing components: compressor with minimum number ofstages and variable stator vanes short annular combustor one-stage supersonic axial turbine withcooled turbine stator and rotor blades made ofnickel-base alloys.As the initial guess for the futuredevelopment the following design pointparameters of core engine are given: polytropiccompressor efficiency c poly 0.879, combustordischarge temperature TET 1650 K, turbinepressure ratio CT PR 4.3, polytropic turbineefficiency e poly 0.85.Fig. 2 demonstrates specific power andspecific fuel consumption (SFC) depending onOPR under given core engine parametersCalculation formulas are well-known can befound in [6]).(near 18:1) variation of SFC is already small,but specific fuel consumption still remains high.It means that 18 looks like a reasonable OPR forthe core engine allowing trade-off betweencomparatively high specific power to confinemass flow-rate and size of compressor and fairlylow SFC.Further arguments invoked in favour ofmoderate value of OPR concern strength-ofmaterial limitation of current technology level.Table 2 shows compressor exit temperatureversus OPR resulted from thermodynamic cyclecalculations. The data are in line with the paper[1] considerations. Fig.3 taken from [1]demonstrates that use of centrifugal compressoras rear stage for OPR larger than 20:1 requireshigh strength materials.Table 2 Compressor discharge temperatureOPR18:1 20:1 22:1 24:126:1T3(K)722 745 767 787806840882921957991T3( F)Specific power (HP/kg/s)450,0440,0430,0420,0410,0400,0390,010 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30OPR0,205SFC (kg/HP/hour)0,2000,195Fig. 3 Compressor exit temperature vs OPR [1].0,1900,1850,1800,1750,17010 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30OPRFig.2 Specific power and SFC vs OPRIt can be seen that maximum specific powerof the engine is achieved at OPR 13:1 andminimum specific fuel consumption is obtainedat OPR 26:1. Note that for moderate OPRThus it can be concluded that a goodestimation of compressor for 9000 HP enginecore is a high rotational speed axial-centrifugalcompressor with minimum (2 1) number ofstages providing moderate values of OPR 18and discharge temperature T3 722 K.More adequately OPR is defined fromconsidering given shaft power (SP) as amaximum work (so called “work potential”)that would be obtained in expanding gas from3
STARTSEV A., FOKIN Yu., STESHAKOV Eu.core exit temperature and pressure to ambientpressure in an imaginary power turbine. Thismethod developed in paper [15] for turbo-shaftengine summarizes the analytical relationshipbetween shaft work delivered by power turbineand irreversibility. According to this approach“compressor is considered as two discrete flowstations wherein the average properties at thecompressor entrance and exit are of interest.Work potential method can be used inconjunction with cycle analysis to estimate totalloss inside the compressor and compare this tolosses in other components such that theperformance of the whole system can beoptimized”. This method relates shaft worklosses (loss in work potential) to flowirreversibilities by examining the entropyincrease in the engine. As a result of [15] “anequation has been obtained which expresses themaximum possible shaft work (named belowshaft power SP) output as a unique combinationof compressor/HP turbine shaft work (namedbelow core shaft power CSP) and burner heat”.This rule is taken from [15] and is presentedhere without derivation: CSP 1 CpT 1 CSPQ 1 CpT 1CpT 1 Sirr / R 1e (1)Fig. 5 demonstrates cooling flow-rates incore turbine. Net cooling flow-rate accounts for16.5% of compressor flow-rate.Fig 5 Distribution of cooling air in coreturbineFrom ambient conditions follows:CpT1 1007 kJ/kg/K 288K 290 kJ/kgFrom given core turbine power and flowrate specific power of core engine can beobtained:CSP 9899.3 kW / 22.3 kg/s 443.9 kJ/kgCSP/CpT1 443. 9 kJ/kg /290 kJ/kg 1.531To evaluate increase of entropy due toirreversibility Sirr it is convenient to userelationship between Sirr and polytropicefficiency poly of compression and expansion. Sirr 1 poly Pexit ln()R polyPinlet(2)If Sirr in compressor is obtained then from(2) follows relationship between compressorpolytropic efficiency c poly and OPR.To evaluate irreversibility Sirr incompressor it is necessary to use as given thedata obtained in a peculiar cycle calculation.Peculiarity of this calculation consists inignoring the turbine cooling, so that value ofTET will be smaller than in real core cycle andequal to 1394K.From given TET and assuming T3 722 Kone can obtain:Q / CpT1 2.338Then from (1) one can obtain: Sirr/R 0.895Design parameters of core turbine are asfollows:CT PR 4.3, e isen 0.865, e poly 0.85Then from (2) one can obtain:( Sirr/R)turbine 0.514
CFD DESIGN AND ANALYSIS OF A COMPACT SINGLE-SPOOL COMPRESSORand( Sirr/R)compressor 0.895 - 0.51 0.385Thus expression (2) is reduced todependence between compressor efficiency ( c)and OPR, if OPR is equal to 18, thencompressor efficiencies should be no less than: c poly 0.879 , c isen 0.8253 Axial-centrifugal compressor designFor this work the following target pressureratios have been chosen: LPC PR 2.6:1 andHPC PR 7:1 to obtain overall PR equal to18:1.To achieve target LPC PR a 2-stage LPChas been scaled from tested prototype (3-stagefan with PR equal to 4.2:1) developed earlier.Scaling coefficient is calculated to match LPCmass flow-rate requested by core. This activityspecifies not only the LPC geometry but alsodesign rotational speed of the axial-centrifugalcompressor.3D RANS performances of the prototypecoincided well with experimental data obtainedearlier approving validity of in-house CFDsoftware. Design tip clearance is 0.35 mm.Geometry and performances of the 2-stage LPCare typical for high tip speed fans (see [7]). LPCrotor and stator blades are low-turning, widechord and of high aspect ratio. Inlet guide vaneis variable (with turning flap). At the outlet ofLPC the flow is axial (excluding near-hubstreamlines). Tip radius of Rotor1 is 211.7 mm.Geometry and gas-dynamics of the 1st LPCstage and 2nd stage are shown in Table 3 and 4.Table 3. LPC. The 1st stage parametersRotor1Stator10.3870.547Hub-to-tip radius ratio1.4260.77Flow Mach numberRotor tip / Stator hub2138Number of blades1.6051.472Solidity0.4720.425Diffusion factor1.7920.983Total pressure ratioTable 4. LPC. The 2nd stage parametersRotor2Stator20.6640.673Hub-to-tip radius ratio1.2360.837Flow Mach numberRotor tip / Stator hub3183Number of blades1.5632.081Solidity0.3980.354Diffusion factor1.5220.978Total pressure ratioLPC design rotational speed is fairly high(Rotor 1 tip speed equals to 487 m/s). Togetherwith moderate value of LPC PR it deliversrather high value of corrected rotational speed tocentrifugal impeller which is good enough toobtain HPC PR 7:1 and to design the impellerin optimal manner. Optimal design of thecentrifugal stage provides high efficiency to thewhole axial-centrifugal compressor. Nowadays,optimal design of a rear (high hub-to-tipdiameter ratio) centrifugal stage ranks as aburning problem (see [8] and [9]).Thus corrected flow-rate, correctedrotational speed and target HPC PR are inputparameters for HPC design. Due to axial flowdischarge by LPC impeller loading is obtainedby its outlet geometry. HPC PR is used todetermine velocity triangle at the outlet ofimpeller. As is known (see [10], [11]), byoptimizing blade loading coupled with highimpeller back-sweep angle out and increasedrelative velocity diffusion ratio Win/Wout it ispossible to increase centrifugal stage efficiencywithout compromise in HPC PR and surgemargin. Large diffusion ratio means increase ofblade height hout at the impeller exit anddiminished meridional velocity Cm out. Increasedblade height offers the advantage of diminishedrelative value of tip clearance. High impellerback-sweep makes more uniform exit flow andwidens range of impeller stable operation.Paper [12] contains valuable formulaeoutlining the flow at the impeller outlet.Explicit formula relating variations of impellerback-sweep angle out and relative velocitydiffusion ratio Win/Wout can be derived underconditions of given (non-varied) HPC PR, tip5
STARTSEV A., FOKIN Yu., STESHAKOV Eu.speed of centrifugal impeller Uout and outletimpeller swirl:Win sin( out ) d(Win/Wout) Cm out d( out) 0 (3)As a result, a back-sweep out 28 , bladeheight hout 20.3 mm and tip speed ofcentrifugal impeller Uout 647 m/s are adoptedto obtain HPC PR 18.Proper attention has to be given to the Uout.Its value has to be limited to control structuraland thermal stress levels and allow currentlyavailable alloy material’s application. Fig.6taken from [13] outlines centrifugal impellermaximum allowable tip speed versus materialtemperature.Fig. 6 Centrifugal impeller maximum allowabletip speed [12].In our case, compressor dischargetemperature is 722 K ( 450 C) andcorresponds to the left boundary of temperatureoperating range shown on Fig.6. As forallowable tip speed, value of 647 m/scorresponds to a titanium-base alloy. Such typetitanium-base alloy is currently available.Specified Uout and known compressorrotational speed give the value of impeller tipradius Rout equal to 280.9 mm. At this step ofimpeller design it is important to specify inducertip diameter. Below inducer tip radius issymbolized by R1s.Paper [14] explains how to choosedimensions and inlet blade angle of impeller fora given flow and pressure ratio. Formula (26) in[14] relates R1s/Rout to the relative flow angle 1sand Mach number M1s at inducer tip diameterand Uout. Recommended by [14] value of 1s is60 . Mach number M1s has to be no more than1.25 to prevent significant shock wave losses.And using formula (26) the value R1s/Rout 0.643 has been obtained, so that R1s 180.6mm.Finally, hub radius of inducer has to bechosen to provide swallowing of givencompressor mass flow-rate, so that inducer hubto-tip radius ratio has been obtained as equal to0.69 completing impeller design.There is else one important questionconcerning intermediate S-shaped duct betweenLPC and HPC. Several trials are required tomatch LPC outlet eye and HPC inlet eye tomake S-shaped duct between LPC and HPC assmall as it possible. For that shroud of the LPC2nd stage has been made descending withcylindrical hub of Stator 2.Completion of the HPC geometry isobtained through outlet system design.Configuration of the centrifugal double-rowbladed diffuser and radial-axial bend isinnovative. Flow deceleration in diffuser islarge, but double-row configuration inhibitsadvent of viscous flow separation, moreover,small total pressure loss is unprecedented. Areacontrolled flow diffusion in radial-axial benddelivers low-speed uniform flow to the inlet ofaxial de-swirl vanes. Requested flow turning inaxial de-swirl vanes is 53 . Static pressure risecoefficient of the outlet system (diffuser deswirl vanes) is equal to 0.84 at the total pressurerecovery coefficient of 0.91.All the HPC design efforts have beensupported by 3D RANS flow simulation with0.4 mm tip clearance in impeller. Applicabilityof in-house CFD software to a centrifugal stageflow simulation has been confirmed in ESPOSAproject (EC FP7) by CFD calculation of highpressure ratio centrifugal compressor of6
CFD DESIGN AND ANALYSIS OF A COMPACT SINGLE-SPOOL COMPRESSORexperimental AI-450S engine developed andtested by IVCHENKO-PROGRESS (Ukraine).Gas-dynamics of the HPC impeller anddouble-row bladed diffuser is shown in Table 5.Table 5. HPC. Impeller and diffuser parametersImpellerDiffuserInlet flow angle57.6 77.3 Inlet flow Mach1.2411.028number (Rotor tip)Outlet flow angle43.2 65.6 Outlet flow Mach0.3820.289number14/1421/21Number of blades7.5520.927Total pressure ratio4 Axial-centrifugal compressor geometryFig. 7 presents compressor dimensions.Axial length of compressor from LPC IGVleading edge to trailing edge of HPC deswirlvanes is 496.8 mm. Axial length of LPCincluding S-channel is 291.6 mm. Axial lengthof HPC impeller is 134.9 mm.Gas-dynamics of the HPC axial de-swirlvanes and combustion chamber pre-diffuser isgiven in Table 6.Table 7. Compressor integral parametersTotal pressureIsentropicratioefficiencyst1 stage1.7610.874nd2 .826compressor5 Axial-centrifugal compressor performancesAfter successful matching of axial LPC andcentrifugal HPC stage there has been made 3Dviscous flow calculation of the compressorperformances for a wide range of RPM toconfirm that surge margin is enough.2018Total pressure ratioAfter LPC and HPC design there has beenmade 3D viscous flow calculation through thewhole axial-centrifugal compressor. As a result,Table 7 presents integral parameters of LPCstages, whole LPC, HPC and whole axialcentrifugal compressor at design point.Fig. 7 Main compressor dimensions.161412108648910 11 12 13 14 15 16 17 18 19 20 21 22 23Corrected flow (kg/s)2018Total pressure ratioTable 6. HPC. De-swirl vanes and pre-diffuserparametersDe-swirl Pre-DiffuserInlet flow angle56.1 3.05 Inlet flow Mach0.2930.167number (Rotor tip)Outlet flow angle3.05 3.24 Outlet flow Mach0.1670.143numberNumber of blades92 Total pressure ncy isentropic0,850,9Fig.8 Axi-centrifugal compressor performances7
STARTSEV A., FOKIN Yu., STESHAKOV Eu.Fig.8presentsCFD-predictedcompressor performances for the range of RPM:100% (n 22025 RPM), 93.5%, 84.8% и 74.8%.Corresponding turning of IGV flap is asfollows: 0 , 0 , 14 , 28 .ConclusionThis paper presents CFD design and studyof compact one-spool 18:1 pressure ratio axialcentrifugal compressor. High isentropicefficiency potential of the compressor (82.5%)is the main novelty of the project and essentiallycaused by: low aerodynamic loading, optimum tipspeed and high flow capacity of LPC optimum design of centrifugal impeller double-row configuration of bladeddiffuser controlled flow diffusion in radial-axialbend.Compressor geometry and its 3D RANSperformances are used as input in the coredesign providing basement for the 9000 HPturbo-shaft engine feasibility.References[1] Welch, G.E., Hathaway, M.D., Skoch, G.J. andSnyder C.A., (2009) “Rotary-Wing RelevantCompressor Aero Research and TechnologyDevelopment Activities at Glenn Research Center”,American Helicopter Society 65th Annual Forum,Grapevine, TX, May 27-29, 2009.[2] K.Rud, (MTU), (2007), “Powerplant for Future HTHHelicopter. Advanced Engine Concepts andTechnologies”, 25th International Helicopter Forum,Sept 19-20, 2007.[3] Goulos, I., et al, (2013), “Rotorcraft Engine CycleOptimization at Mission Level”, Proceedings ofASME Turbo Expo 2013, GT2013-95678.[4] Kroger, G., Siller, U., Moser, T., and Hediger S.,(2014), “Towards a Highly Efficient Small ScaleTurboshaft Engine. Part I: Engine Concept andCompressor Design”, Proceedings of ASME TurboExpo 2014, GT2014 – 26368[5] Stephen A. Anderson, Ronald Trumper and GaryWeir, (PWC), (2010), “Hybrid Compressor”, UnitedStates Patent Application Publication No. US2010/0232953 A1.[6] Kurzke, J., (2007), “ About Simplifications in GasTurbine Performance Calculations”, Proceedings ofASME Turbo Expo 2007, GT2007-27620[7] H.E.Messenger and E.E.Kennedy, (1972), “TwoStage Fan. Aerodynamic and Mechanical Design”,NASA CR-120859.[8] Lurie, E.A., et al., (2011), “Design of a HighEfficiency Compact Centrifugal Compressor”,American Helicopter Society 67th Annular Forum,May 3-5, 2011.[9] Ralf von der Bank, et al, (2014), “LEMCOTEC –Improving the Core-Engine Thermal Efficiency”,Proceedings of ASME Turbo Expo 2014, GT2014 –25040.[10] Mileshin, V.I., Startsev, A.N., and Orekhov, I.K.,(2003), "CFD Design of a 8:1 Pressure RatioCentrifugal Compressor", Proceedings of IGTC 2003Tokyo,Japan,November2-7,2003,IGTC2003Tokyo TS-043.[11] Takanori Shibata, et al, (2009), ”PerformanceImprovement of a Centrifugal Compressor Stage byIncreasing Degree of Reaction and Optimizing BladeLoading of a 3D-Impeller”, Proceedings of ASMETurbo Expo 2009, GT2009 – 59588[12] Shum, Y.K.P., Tan, C.S. and Cumpsty N.A., (2000),“Impeller-Diffuser Interaction in a CentrifugalCompressor”, Journal of Turbomachinery, October2000, Vol. 122, pp. 777-786[13] Singh, B., (1991), “Small Engine Component Study”,NASA CR-175079, Teledyne CAE Report No 2224[14] Daniel Rusch and Michael Casey, (2012), “TheDesign Space Boundaries for High Flow CapacityCentrifugal Compressors”, Proceedings of ASMETurbo Expo 2012, GT2012-68105[15] Christofer D. Wilson, David W. Riggins, Bryce Rothand Robert McDonald, (2002), “PerformanceCharacterization of Turboshaft Engines: WorkPotential and Second-Law Analysis”, AmericanHelicopter Society 58th Annular Forum, June 11-13,2002.Copyright StatementThe authors confirm that they, and/or their company ororganization, hold copyright on all of the original materialincluded in this paper. The authors also confirm that theyhave obtained permission, from the copyright holder ofany third party material included in this paper, to publishit as part of their paper. The authors confirm that theygive permission, or have obtained permission from thecopyright holder of this paper, for the publication anddistribution of this paper as part of the ICAS 2014proceedings or as individual off-prints from theproceedings.8
are reduced to dependence between compressor efficiency ( c) and OPR, so that optimum OPR is 18, if 82.5% compressor isentropic efficiency could be achieved. The second part of the paper outlines detailed CFD design of the compressor. This one-spool high-speed compressor consists of a two-stage low pressure axial compressor (LPC)
refrigerator & freezer . service manual (cfd units) model: cfd-1rr . cfd-2rr . cfd-3rr . cfd-1ff . cfd-2ff . cfd-3ff . 1 table of contents
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A.2 Initial Interactive CFD Analysis Figure 2: Initial CFD. Our forward trained network provides a spatial CFD analysis prediction within a few seconds and is visualised in our CAD software. A.3 Thresholded and Modified CFD Analysis Figure 3: Threshold. The CFD is thresholded to localise on
CFD Analysis Process 1. Formulate the Flow Problem 2. Model the Geometry 3. Model the Flow (Computational) Domain 4. Generate the Grid 5. Specify the Boundary Conditions 6. Specify the Initial Conditions 7. Set up the CFD Simulation 8. Conduct the CFD Simulation 9. Examine and Process the CFD Results 10. F
performing CFD for the past 16 years and is familiar with most commercial CFD packages. Sean is the lead author for the tutorial and is responsible for the following sections: General Procedures for CFD Analyses Modeling Turbulence Example 3 - CFD Analysis
EFD, CFD, and UA laboratories. EFD labs were improved and UA was introduced. Complementary CFD labs were also introduced using an advanced research code modified for limited user options. From 1999 to 2002, the research CFD code was replaced by the commercial CFD software (FLUENT) and refinements were made and
The CFD software used i s Fluent 5.5. Comparison between the predicted and simulated airflow rate is suggested as a validation method of the implemented CFD code, while the common practice is to compare CFD outputs to wind tunnel or full-scale . Both implemented CFD and Network models are briefly explained below. This followed by the .
Emphasis is on comparing CFD results, not comparison to experiment CFD Solvers: BCFD, CFD , GGNS Grids: JAXA (D), ANSA (E), VGRID (C) Turbulence Models: Spalart-Allmaras (SA), SA-QCR, SA-RC-QCR Principal results: Different CFD codes on same/similar meshes with same turbulence model generate similar results
