super computing applications in internal combustion engine design and

8/3/2019 Super Computing Applications in Internal Combustion Engine Design And

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Supercomputing Applications in Internal Combustion Engine Design andAnalysis

Woo Tae Kim and Kan g Y. HuhMechanical Engineering D epartmentPohang University of Science and TechnologySan 3 1 Hoyjadong, Pohang, Kyung buk, Koreakw t @ [email protected]

AbstractSupercomputing applications in intema l combustion en-

gine design and an alysis ure introduced in this papel: Thegrid generation procedure, f low f ield modeling, fuel sprayand combustion models are briefly reviewed. Results f o rthree example cases, motoring flow in a pent-roof en-gine, charge distribution in a DISC engine, and CoherentFlamelet combustion model are given to illustrate the cur-rent status and future prospects of supercomputing in theautomotive industry. I t is expected that fur the r develop-ment and cost reduction in supercomputing will allow ac-curate simulation with an increased number of grid pointsand more realistic physical m odels.

1. IntroductionIn-cylinder dynamics of internal combustion engines in-

volves a num ber of complex, closely coupled physical andchemical processes. They include transient turbulent flowsinteracting with fuel sprays, heat transfer and chemical re-actions involving ignition, combustion, knocking and pol-lutant production. Du e to recent developments in internalcombustion en gine modeling, computational fluid dynam-ics, and supercomputing, it has become possible to performa more realistic simulation of the in-cylinder phenomena.Applicable areas of numerical simulation are expanding toprediction of major engine parameters and optimization ofthe existing and new engine designs.

The numerical analysis procedure of internal combus-tion engines may be divided into three steps. The first stepinvolves grid generation and proble m definition. A gridis a system of points where the physical variables are de-fined and calculated. Grid generation is usually the mosttime-consuming part for the analyses with complex geom e-

tries. Sometim es it may be necessary to comprom ise thegeometry to a certain extent to avoid an unacceptable grid.Proble m definition consists of determination of the primaryvariables and assumptions with specification of initial andboundary conditions. The second is the code execution stepwhich generates results through tuning of various parame-ters such as the time step or arbitrary mode l constants. Ifthe calculation involves a new physical process, it is nec-essary to modify the existing models or to develop a newmodel, which must be followed by validation of the newimplemented model. The last step is the post-processingwhich extracts useful information from the calculation re-sults. Since post-processing has a great influence on thequality of final conclusions, selection and dev elopm ent of aproper post-processing method is very important.

There are several engine analysis codes in current usee.g. the KIVA code series[ 1,2] by Los Al amo s NationalLaboratory, EPISO and SP EED code[3,4] by Imperial Col-lege of Sci. & Tech. and the c omme rcial softwares such asSTAR -CD and FIRE. In this paper, we introduced the ma-jor physical models and numerical procedure for in-cylinderprocesses in a modified version of KIVA-11, which are ap-plied to the three application calculations in the followingsections.2. Grid generation

Since an internal combustion engine usually involves acomplex three-dimensional domain, the c ommercial pack-ages with enhanced geometry modeling capabilities arecomm only used. In this study, the commercial softwareICEM-C FD/CAE [S] is used for grid generation accordingto the following steps.

(1) Transform the IGES format CAD data into the ICEMDDN data format.

( 2 ) Design the block structure which efficiently divides

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the com putational domainfaces.

(3) Define the subfaces which comprise the domain sur-(4) Define the dom ains which make the whole geometry.( 5 ) Specify the node number in the associated main(6) Generate the node points.(7) Inspect the quality of the generate d grid.(8 ) Modify th e grid if necessary.(9) Generate a grid input file for the analysis code.

edges.

Step (2) is the mo st important step because the block struc-ture determin es the overall grid quality. An improperlydesig ned block structure results in severe distortion andstretching of meshes which may restrict the time step to avery small value. Exam ples of three-dimensional CAD dataand generated grids are show n in Figure 1 an d Figure 2.

\ \ \

/ .\ . \\ /

\ /\ /

\ /\ /.,

Figure 1. Engine CAD data

3. Flow field modelingGoverning equations for flow field modeling consist of

the mass, m omentum , energy, and IC -& turbulence equations.A turbulence model is required because it is impossible toresolve the wide range of spatial and temporal scales of tur-bulent eddies. Auxiliary constitutive equations and experi-mental correlations are also needed for closure of the givenequations. R elevant source terms may be ad ded to the flowequations for the spray and combu stion phenom ena.

The continuity equation for the mt h species is(1)P ,-- +v . (p,Z) = v [ p D O ( P ) ]at P

Figure 2. Computational grids

where p, is the density of the mth species, p is the totalgas density, D is the Fick s law diffusion coefficient, and Uis the flow velocity. By sum ming up the species continu-ity equations, the following mass conservation equation isobtained.

- v (p.) = 0d tThe mom entum equat ion is as follows.

where p is the fluid pressure, a is a dimensionless quan-tity associated with the Pressure Gradient Scaling(PGS)method[6], and 9 is the specific body force. The viscousstress tenso r is Newtonian in the form :

where p an d X are the first and second coefficients of vis-cosity.

The internal energy equation isv . i f P E ( 5 )

where I is the specific internal energy. T he heat flux vectorJ is the sum of contributions due to heat conduction andenthalpy diffusion:

+J = -7Tl

where T is the fluid temperature and h, is the specific en-thalpy of the mt h species.

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The turbulence equations are to make reliable prediction of the in-cylinder phenomen arelated with fuel sprays.

dPk- v ' ( pu ' k )=at2 P- - p k V . d+ g : Vu'+Q . [(---)Vk] - E (7)3 p r k

2 P-(p , , ) p e V . U '+ V ' [(-)VelPrEwhere k is the turbulent kinetic energy and E is its dissipa-tion rate. The quantities e,,, eE2,c,,, P T k , and PrE areconstants whose values are determined from experimentsand theoretical considerations.

4. Spray modelingFuel sprays have been a major research topic because

the fuel spray behavior determines the com bustion charac-teristics of diesel engines. Th ey also play an impo rtant rolein port-injection spark-ignition engin es and direct-injectionstratified charge(D1SC) engines. A fuel spray is numeri-cally represented by the droplets of 3000 - 5000 parcelswhich share the same condtions in each parcel - ass,temperature, and mean velocity, etc. - nd tracing themby a stochastic Lagrangian method. The word 'stochastic'means that sampling procedure is performed randomly fromthe assumed probability density distribut ions which governthe droplet properties at injection timing and subsequentdroplet behaviors. The fuel injection process is modeled bycreating new spray parcels with the pres cribed initial cond i-tions. Interaction with the turbulent flow field, droplet evap-oration, breakup, collisonk oalesce nce, and spray-wall im-pingement, etc. may be easily taken into account for eachLagrangian parcel.

In the following calculations, the breakup and col-lisionkoalescence models are excluded with the initialdroplet radius determ ined by a probability density func-tion about the measured sauter mean radius. It is to re-move the large uncertainties associated with these models.In the modified version of KIVA-11, a new spray atomiza-tion model is implemented to consider the effect of tur-bulent fluctuations in the nozzle exit flow as well as wavegrowth instability by gas inertia force. Also a spray-wallimpingement model based on single droplet experimentsis implemented to provide more accurate information oncharge distribution. In som e engine types, spray-wall orspray-piston collision is induced intentionally to achieve adesirable mixing pattern of fuel and ambie nt air. Furthervalidation against extensive spray experiments are needed

5. Combustion modelingVigorous researches are going on in combustion model-

ing, altho ugh it is still in its initial stage in practical applica-tions. Available turbulent combustion models are the eddybreakup(EBU ) model[7], assumed probability density func-tion(PDF) model[8] and cohe rent flamelet model(CFM)[9],etc. The CFM views the turbulent flame as a collection oflamin ar flamelet element s emb edde d in turbulent flow. Itis applicable to both premixed and nonpremixed flame onthe basis of the lam inar flamelet concept. I ts important ad-vantage is deco upled treatment of the chem ical reaction andturbulent flow.The current form of CFM is introduced in the following.The ma ss burning rate per unit volume is given as

where pv is the fresh gas density, U L is the laminar flamespeed, and C is the flame density which is the flame areaper unit volume. The product IOUL epresents the strainedlaminar flame speed, where IO in the mean stretch fac-tor. Chem istry and m olecular effects are represented by U Lwhile turbulence effects are represented by C .Th e conservation equa tion for the flame density C is

ax a- - (UiC) =at axiuL P ( 1 0 )d vt dC U' Y(--) + a--C - paxa axi l t , (ZIP- E)

where cy, p are constants( cy=2.0, p=O.l),U' is turbulenceintensity, l t , is a constant length scale( ltC=0.126cm).is the mean fuel mass fraction, YO s the fuel mass fractionin the fresh mixture. The sec ond term on the right handside is production of flame area by turbulent strain and thethird term is flam e annihilation by mutual collision of flamesurfaces.6. Applications6.1. Intake and compression flows in a motoring

pent-roof engineComplicated three dimensional geometry with moving

boundaries such as valves has been the major difficultyamong CFD practitioners. It is no trivial task to generatea suitable structured or u nstructured grid for the domain in-cluding the port/valve/cylinder/piston applicable to the en-tire four stroke cycle. In this study a simple technique is

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developed to hand le the valves in an Eulerian doma in. Itdoes not require generation of a body-fitted grid and takesa comparable CPU time as the case with no obstacle in thesam e grid.

Table 1. Engine specifications of a motoringpent-roof engine

BoreStrokeCom pression ratioDisplacem ent volumeEngine speedIntak e pressureIV OIV CValve lift

86.0 mm86.0 mm

9 .21998 cc1000

1 O atm10deg. BTDC55 deg. ABDC

9.0 mm

The engine has a four valve pentroof head with centralspark plug and siamesed intake ports. The eng ine specifica-tions are summa rized in Table 1. A half portion of the en-gine is modeled du e to symmetry and only the intake valveand port are modeled excluding the exhaust side. Calcu-lations are carried out through the intake and compressionstroke, starting shortly after intake valve opening and end-ing at the top dead cen ter(TDC) at 360 deg. CA . Initially thewhole domain is assumed to be quiescent at the atmosphe ricpressure and the piston moves at 1 000 rpm.

Bec ause the in-cylinder fluid motion is very comp lex andthree dimensional, it is difficult to describe the completeflow structure. Figu re 9 shows the velocity fields at thevalve center plane. At 6 0 deg. CA a downward motion pre-vails in the cylin der and valve tip vortices are forme d aroundthe valve periphery. T he flow in the port is very strong withthe maximum velocity of about 30 - 40 m/s. The strongintake flow is still domin ant at 1 20 deg. CA . The reciculat-ing motion on the exhaust valve side is stronger than thaton the intake valve side. On the overall a tumbling m otionis formed in the cylinder but restricted due to the down-ward motion of intake air along the intake side wall. At240 deg. CA deformation of the organized tumble motionoccurs. T he upward motion of piston governs the flow andcontinues to 300 deg. C A. In the later stage of the compres-sion stroke , squish flow occurs into the head volume.

Figure 3and 4 show variation of the tumble ratio a ndthe mass averaged turbulent kinetic energy. The tumble ra-tio is a dimensionless m easure of the angular mom entum ofthe in-cylinder fluid about the axis which is perpendicularto the valve cente r plane. The tumble ratio increases by thestrong intake flow and reaches the maxim um at ab out 120deg. CA . Then it decreases due to weak intake flow andthe down ward motion of the piston until intake valve clos-

-0 1 4

Figure 3. Tumble ratio about x-axis7E5

Crank AngIe(deg.)

Figure 4. Mass averaged TKE variation

ing(1VC). It increases again d ue to angular mom entum c on-servation mechanism and then decreases as the compressioncauses the whole structure to break dow n. The m ass aver-aged turbulent kinetic ene rgy(T KE) maintain s its level afterZVC. The T KE carried into the cylinder through the intakevalve decays fast and does not contribute to the TK E duringthe com pression stroke.6.2. Charge distribution in a lean burn DISC engine

Extensive research works have been perform ed to estab-lish stable combustion of a DISC engine by optimizing thefuel/air mixture formation process[ 10,11,12]. The co mbus-tion chara cteristic s of a gasoline DISC engine may be di-vided int o two different regim es by the injection timing. Ina full load condi tion, fuel is injected durin g the intake strokeand forms a homoge neous mixture with air in the combus-tion chamber. In this case the combustion characteristicsis similar to that of a port injection engine. In a part loadcondition, on the other hand, fuel injection occurs in thelater stage of the compression stroke, to form a stratifiedcha rge around the spark plug at ignition timing. It is to re-alize leaner combustion with improvem ents in the thermalefficiency and NOx emissions.The test engine is a 4-valve 4-stroke gasoline DI en-gine with a pent-roof type combustion chamber. The en ginespecifications and operating conditions are given in Table 2 .

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BoreStrokeIV OIV CEVOEV CInjection durationTotal amount of fuelSpray anglelwidthNozzle radius

Engine speed

Table 3. Notationsof a DISCengine computa-tion

83.0 mm90.5 mm1500

6 deg. BTDC46 deg. ABD C130 deg. ATDC10deg. ATDC

9 deg.10.6642mglcycle60 deg. 16.05 8 deg.

0.8 mm

type1Figure 5center inj.

For a full load condition; fuel injection starts early duringthe intake stroke, from around TDC to 150 deg. after topdead center(ATDC); initial spray dispersion is m uch higherthan that of the late injection case at a part load co ndition.It is obvious that there can be more causes that may leadto bad operation with the late injection case su ch as mis-fire or incom plete combustion du e to mal-distribution of themixture. Thus, we focused on analysis of the late injectioncase at a part load condition. T he ranges of the parametricstudy and so me notations are given in Table 3an d Figure 5 .

Two different injection positions are considered here,center and corner injection, with a tilted injector in bothcases(Figure 5). As shown in Figure 10,center injection ispreferable in some repect for it assures existence of a richmixture near the spark plug. But possible impingem ent ofthe liquid spray on the spark plug may require a high energyignition system or other ignition technologies. Th e cornerinjection case results in a poor charge distribution aroundthe spark plug with a flat piston, although a high pressureswirl injector is employed to enhance mixing. To improvemixture distribution in the corner injection case, a curvedpiston top geometry is suggested. It consists of a pistonbowl on the injection side to make the impinging fuel di-rected toward the spark plug. The equivalence ratio andother results at 40 deg. CA after the end of injection incase 3 are shown in Figure 1O(c), which are easily com paredwith the results of Figure 10(b). Spray-wall impingment is

type2ltype2p piston geom etryFigure 5 with bowl)cor ner in;. (tYPe2P

Figure 5. Injection position

retarded as the spray is shifted upward du e to the mean flowinduced by the piston bowl. A stable vortex formed in thevertical direction drives the spray toward the center of th ecombustion chamber and enhances fuel-air mixing.

6.3. Spark-ignited turbulent premixed flame in aclosed vesselIn this study the CFM introduced in Section 5 is ap-plied to ignition and flame propagation in an enclos ed cubicchamber. A modeling strategy for the ignition, laminar andturbulent propagation phase is proposed and applied with an

appropriate switching criterion. Th e same form of the trans-port equation is applied with the total rate of strain givenas the sum of the laminar and turbulent rate of strain. Thisstrategy is to guarantee a calculation procedure with sm oothtransition between phases. This strategy can be directly ap-plied to a real internal combustion engine simulation whichincludes the ignition, flame propagation, and flame qu ench-ing at the wall.

The shape of the flame when the flame quenches at thewall is shown in Figure 6.Figure 7 shows distribution of theburned mass fraction and flame density as the flame touche sthe wall. Figure 8 is the dimensionless turbulent burningvelocity vs. the turbulent intensity. T he calculated turbulentburning velocity agree s qualitatively with the correlation.

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Z

X Y

Flame density1 4 2 8 9 21 0 7 1 6 90 7 1 4 4 5 80 357229

Figure 6. Flame quenching at the wall

7. SummaryRecen t progress in sup ercomp uting capabilities has al-

lowed more realistic simulation of the in-cylinder phenom -ena of internal combustion engines such as com plex three-dimensional g eometry, turbulent flow field, fuel spray, com -bustion and heat transfer. Thre e exam ples of applicatio ncalculations performed on CRAY -YMP C90 are presentedin this paper to show the current status and future prospectsfor supercom puting in internal combustion engin e designand analysis. Use of supercom puting in comp utational fluiddynam ic analysis is rapidly spreading in various engineer-ing disciplines, one of which is the automotive industry. It isexpected that further developm ent and cost reduction in su-percom puting will allow an increase in the total nu mber ofgrid points and implementation of more com plicated physi-cal models to obtain mo re accurate simulation results.Acknowledgement

The financial support for this work from Systems Engi-neering Research Institute (SERI) and Advan ced Fluid En-gineering Research Center (AFER C) funded by Korea Sci-ence and Engineering Foundation (KO SEF) is appreciated.References[ I ] A. A. Amsden, J. D. Ramshaw, P. J. ORourke, and J. K.Dukowicz, KIVA: A Computer Program fot Two- and Three-Dimensional Fluid Flows with Chemical Reaciions and FuelSprays, Los Alamos National Laboratory report LA- 10245-MS,Feb. 1985.[2] A. A. Amsden, P. J. ORourke, and T. D. Butler, KIVA-II : A Computer Program fo r Chem ically Reactive Flows withSprays, Los Alamos National Laboratory report LA-1 1560-M S,

May 1989.[3] A. D. Gosman, K. Y. Huh, B. S . Tabrizi, and Q. Zhang, Th eEPISO-SPRAY Computer Code fo r Prediction of Fuel Spray aridAir Motion in Motored Internal Combustion Engines, Manual forEPISO-SPRAY Code, Dec. 1 987.[4] SPEED Manual, Version 2.200, 1994.[SI ICEM CF DK AE Users Manual, October, 1993.[6] J. D. Ram shaw, P. J. ORo urke and L.R. Stein, Pressure Gradi-ent Scaling Method fo r Fluid Flow with Nearly Uniform Pressure,J. Comput. Phys. 58, p.361, 1985171 D. B. S palding, Thirteenth Sym posium (Internation al) onCombustion, The Combustion Institute, p.649, 1971.[8] R. Borghi, Prog. Energy C ombust. Sci. 14:245, 1988.[9] E. Maistret, N. Darabiha, T. Poinsot, D. Veynante,E Lucas, S.Candel, and E. Esposito, in Numerical Combustion (A. Dervieuxand B. Larrouturou Eds.), Springer-Verlag, Berlin, p.98, 1989.[IO] Seiko Kono, Development of the Siraiijied C harge and StableCombustion Method in DI Gasoline Engines, SAE 950688, 1995.[ l l ] T. Kume, Y. Iwamoto, K. Iida, M . Murakami, K . Akishino,and H. Ando, Combustion Coriirol Technologies fo r Direci ln jec-tion SI Engine, SAE paper, 1996.[I21 K. Kuwahara et al., Opiimization of In-cylinder Flow aridMixing f o r a Center-spark Four-valve Engine Employing the Corz-cept of B arrel-stratification, SAE 940986.

(More figures in the following pages)

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BMF0.9081 30.8173170.7265040.6356910.5448780.4540650.3632520.2724390.1 8162 60.09081 3

A987654321

4907654321

c37.684133.91 5730.1 47326.378922.610518.842115.073611.30527.536823.76841

Figure 9. Velocity fields at the valve centerand streamlines

Figure 7. Distributions of the burned massfraction and flame density

UIU,

(a) 120deg. CA

Y

(b) 60 deg. CA

,-.

(d ) 240 deg. CA

( c ) 120deg. CA

(e) 300 deg. CA

Figure8. Dimensionless urbulent burning ve-locity vs. the turbulence intensity

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equivalence atio (injection point-fixed y)typet case3CA 310deg~~ -. . _

equivalence atio (injection point-fixed y). . - .__ype2sase3CA3tOd~__ equivalence atio (injection point-fixed y)- typezp case3 CA 3todegj

Level 1 2 3 4 Level 1equvrat 0223486 0670339 1 11719 156404 eqwrat 0 3 0 9 1 5equivalence ratio (injection point fixed y)type2pcase3 CA 320degequivalence atio (injection point fixed y)type1 case3 CA 320d89

equivalence atio (injection point fixed y)type2 case3 CA 320deg~~~~

I I

equivalence atio (iniectionpoint4md y)- type1 case3 CA 330deg . equivalence atio (inpction pomt-fixed y)type2 case3 CA 330deg.. . . . ~~~~~ equivalence ratio (injectionpoint fixed y)type2p case3 CA 330degl

velocllylteldal CA310degtypet case3 (40deg alter in! end)- - ~- ~~~

velocity ield at CA 3tCdegtype2case3 (40deg aner in1 end)- _ ~ _ _ -

velocity ield at CA 3tCdegtypezp case3 (40deg after n1 end) 1

Level 1 2 3 4 Level 1 2 3 4 ~ Level 1tke 265707 65tt7.5 tot744 142331 1ke. 269396 645301 102121 139711 tke. 2 5 3 4 5 9 609536 96561.3 132169 Itke (injectionpoint-fixed y)typet case3 CA 3lCd-q tke (mjectionpoint-fixed y)- type2~_ ~~case3CA 310d.g

Figure 10. Equivalence ratios, velocity field, and TKE distribution

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super computing applications in internal combustion engine design and

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