International journal of automotive technology, tập 11, số 5, 2010

CONCLUSIONS 1 The Temperature Phase model for soot calculation was applied to simulate a wide range of part load operating conditions, varying injection timing, EGR-rate and injec-tion p

International Journal of Automotive Technology , Vol 11, No 5, pp 611 − 616 (2010)

611

ENHANCEMENT OF NOx-PM TRADE-OFF IN A DIESEL ENGINE

ADOPTING BIO-ETHANOL AND EGR

S JUNG 1)* , M ISHIDA 2) , S YAMAMOTO 2) , H UEKI 3) and D SAKAGUCHI 2)

1)Research and Development Department, Daihatsu Diesel Mfg Co Ltd., Shiga 524-0044, Japan

2)Graduate School of Science and Technology, Nagasaki University, Nagasaki 852-8521, Japan

3)Faculty of Engineering, Nagasaki University, Nagasaki 852-8521, Japan(Received 12 December 2009; Revised 19 February 2010)

ABSTRACT− For realizing a premixed charge compression ignition (PCCI) engine, the effects of bio-ethanol blend oil and exhaust gas recirculation (EGR) on PM-NOx trade-off have been investigated in a single cylinder direct injection diesel engine with the compression ratio of 17.8 In the present experiment, the ethanol blend ratio and the EGR ratio were varied focusing

on ignition delay, premixed combustion, diffusive combustion, smoke, NOx and the thermal efficiency Very low levels of 1.5 [g/kWh] NOx and 0.02 [g/kWh] PM, which is close to the 2009 emission standards imposed on heavy duty diesel engines in Japan, were achieved without deterioration of the thermal efficiency in the PCCI engine operated with the 50% ethanol blend fuel and the EGR ratio of 0.2 It is found that this improvement can be achieved by formation of the premixed charge condition resulting from a longer ignition delay A marked increase in ignition delay is due to blending ethanol with low cetane number and large latent heat, and due to lowering in-cylinder gas temperature on compression stroke based on the EGR It is noticed that smoke can be reduced even by increasing the EGR ratio under a highly premixed condition.

KEY WORDS : PCCI engine, Bio-ethanol, EGR, PM-NOx trade-off

1 INTRODUCTION

Recent researches aiming at almost zero emissions on NOx

and PM in diesel engines have shown that a homogeneous

charge compression ignition (HCCI) engine or PCCI

engine will be a promising way to accomplish the target In

order to achieve the severe emission standards imposed on

diesel engines such as the post new-long term emission

standards for 2009 in Japan, the EURO VI standards for

2012 and the US standards for 2010, it is necessary not

only to depend upon the aftertreatment system with NOx

catalyst and PM catalyst but also to suppress NOx and PM

formations in spray combustion process by forming the

PCCI or HCCI condition

The HCCI engine showed some problems on the limited

operation range due to misfire and knock (Jung et al., 2005;

Ishida et al., 2006, 2007, 2008; Jung et al., 2007), then, the

PCCI engine is thought to be better than the HCCI engine

in order to expand the engine load range In the recent

studies on a PCCI engine operated with a high EGR and

low cetane number fuel (Ogawa et al., 2007), and Li et al

(2007, 2008) showed that smoke emission decreased with a

longer ignition delay due to the low temperature combustion

under the ultra-high EGR condition, and smoke was

dependent strongly upon the premixing time from the end

of fuel injection to the onset of ignition on the basis ofcombustion tests with large quantities of cooled EGR andlow cetane number fuels On the other hand, bio-ethanol as

a carbon neutral fuel is one of the alternative fuels, and it iseffective to reduce carbon dioxide emissions Kamio et al.(2007) investigated the effect of ethanol fuels on HCCI-SIhybrid combustion using dual fuel injection Ishida et al.(2004) showed the effectiveness of gasoline blended withgas oil in a diesel engine experimentally In this case,smoke was reduced markedly by blending gasoline havinglow cetane number and low evaporation temperature,resulting from a longer ignition delay

The objective of the present study is to show one of theapproaches for achieving low NOx and low PM com-bustion in a diesel engine by utilizing bio-ethanol The gasoil was blended with bio-ethanol for realizing a PCCIengine in the present experiment The effects of ethanolblend ratio and EGR ratio on ignition delay, premixed com-bustion, diffusive combustion, smoke density, concentrations

of NOx, unburned hydrocarbon and carbon monoxide, andthe thermal efficiency were investigated in detail

2 EXPERIMENTAL APPARATUSThe test engine is a single cylinder high-speed naturallyaspirated direct injection diesel engine, the type NFD 170-(E) manufactured by YANMAR Co., Ltd The bore is 102

*Corresponding author. e-mail: sukho1001@hotmail.com

612 S JUNG et al.

mm, the stroke is 105 mm, and the compression ratio is

ε=17.8 Test fuels are gas oil having cetane number of 55

and ethanol with that of about 8 Table 1 shows properties

of tested fuels The blend fuel named EtOH30 consists of

68 vol% gas oil, 29 vol% ethanol with 3 vol% octanol as a

surface-active agent, and EtOH50 consists of 48% gas oil,

48% ethanol and 4% octanol The lower heating values of

gas oil and ethanol are 42.9 and 26.8 MJ/kg respectively

Ethanol shows a lower cetane number, lower evaporation

temperature and larger latent heat compared with gas oil

Figure 1 shows the experimental test system The

com-bustion tests were carried out under the conditions of a

constant engine speed of 1,200±5 rpm, a constant suction

air pressure of 0.1013 MPa at the intake manifold and a

constant intake temperature of TIN=40oC even in the case

with EGR, in other words, the cooled EGR EGR gas was

charged into the mixing chamber located at 1,800 mm

upstream of the intake manifold, and the fuel was injected

into the combustion chamber directly at a constant

injec-tion timing of 5o CA BTDC CO2 concentration% wasmeasured at both intake and exhaust manifolds respectively

to calculate the EGR ratio XEGR

In the exhaust gas analysis, exhaust gas temperature Te

oC, concentrations of carbon monoxide CO ppm, totalunburned hydrocarbon THC ppm and nitrogen oxides NOxppm, and smoke density were measured respectively asshown in Figure 3 later The time-history of in-cylinderpressure was measured using the piezo type sensor and thisoutput was sampled every one-fourth degree in crank angle

by means of the 4 channel combustion analyzer CB-467manufactured by Ono Sokki Co Ltd The time-history ofcombustion pressure was the ensemble average sampledover continuous 350 engine cycles The data were trans-mitted to the personal computer and recorded on harddisks

3 RESULT AND DISCUSSION3.1 Effects of Ethanol Blend Ratio and EGR Ratio onCombustion Time-history and Engine PerformanceFigure 2 shows a change in time-history of combustion due

to ethanol blend and EGR as well under the high engineload of Pme=0.51 MPa CA of the abscissa denotes thecrank angle degree, and P, dQ/dθ and Lift in the ordinatedenote a measured in-cylinder pressure, apparent heatrelease rate and needle valve lift respectively In the experi-ment, the EGR ratio was increased while the intake temper-ature was kept constant As the ethanol blend ratio increasesand also the EGR ratio increases, ignition timing is retard-

ed markedly, then, the premixed combustion becomeslarger and the diffusive combustion decreases remarkably.Figure 3 shows changes in exhaust emissions and fuel

Table 1 Properties of test fuels

Gas oil EtOH30 EtOH50

Figure 1 Combustion test system

Figure 2 Effects of ethanol and EGR on combustion history (Pme=0.51 MPa)

time-ENHANCEMENT OF NOx-PM TRADE-OFF IN A DIESEL ENGINE ADOPTING BIO-ETHANOL AND EGR 613

consumption due to engine load, where the parameters are

fuels of gas oil and 50% ethanol blend fuel (EtOH50) and

the EGR ratio The brake specific fuel consumptions reduced

by lower heating value of each fuel are almost the same

between these three cases The one of EtOH50 is a little

lower than that of gas oil alone in the high load range NOx

increases by means of ethanol blending because the

pre-mixed combustion increases due to longer ignition delay

and due to lower evaporation temperature of ethanol, on

the other hand, it is reduced markedly by EGR as shown in

Figure 3 Smoke is reduced remarkably in the case of 50%

ethanol blend fuel, in addition, it is decreased by EGR

further

Figure 4 shows smoke change rate due to EGR ratio

Smoke/Smoke0 in the ordinate denotes a proportion of

smoke without EGR (Smoke0) to smoke with EGR in eachcondition It should be noticed that, only in the case ofEtOH50, smoke was reduced by increasing the EGR ratio.The reason why smoke is reduced by EGR will be clearlyshown in the latter section of this paper

3.2 Ignition Delay Due to Ethanol and EGRFigure 5 shows the definitions of a ignition timing, ignitiondelay and approximated diffusive combustion curve of theheat release rate The ignition timing was defined as a zero-cross point of the dotted line tangential to the heat releaserate curve in a initial premixed combustion stage Theignition delay was defined as the crank angles between thestart of injection and the ignition point The heat releaserate curve during the diffusive combustion period wasapproximated by the Wiebe’s function (Miyamoto et al.,1985) as shown in Figure 5, and the cumulative heat release

of diffusive combustion Qd was calculated by integratingthe Wiebe’s function

Figure 6 shows change in the ignition delay due to theEGR ratio; where parameters are fuel and engine load Theignition delay is dependent largely upon the fuel cetanenumber; cetane numbers of the fuels vary from 55 of gasoil to 27 of the fuel EtOH30 and 18 of the fuel EtOH50 In

Figure 3 Effects of ethanol and EGR on exhaust emissions

and engine performance

Figure 4 Change in smoke due to EGR

Figure 5 Definitions of ignition timing, ignition delay anddiffusive combustion

Figure 6 Effects of ethanol and EGR on ignition delay

614 S JUNG et al.

the ethanol blend fuels, ignition delay is increased mainly

by the lower cetane number and secondarily by the larger

latent heat The smaller the cetane number is, the longer the

ignition delay is The ignition delay increases markedly

with increase in the ethanol blend ratio, and also increases

further by increasing the EGR ratio However, variation of

ignition delay due to engine load is relatively small except

for the cases with the high EGR ratio Increase in ignition

delay due to EGR is based on lowering the in-cylinder gas

temperature on compression stroke The longer ignition

delay promotes premixing between the fuel and intake

charge

Figure 7 shows change in the cumulative heat release

during diffusive combustion (Qd) due to ignition delay,

where the parameter is the engine load The data in the

figure include the cases with different ethanol blend ratio

and the cases with different EGR ratio Qd, which seems to

be a main factor of smoke emission, decreases almost

linearly with increase in ignition delay at any engine load

although Qd is larger at higher load because of larger fuel

injection quantity

3.3 Relationship between PM and Diffusive Combustion

The mass rate of diffusive combustion md was calculated

by the following equation;

(1)where Hu is the lower heating value of the fuel

Assuming that higher evaporation temperature

compo-nents of gas oil burns in the diffusive combustion process,

the lower heating value of gas oil is applied for Hu because

ethanol has fairly lower evaporation temperature compared

with gas oil, and it burns in the premixed combustion stage

Figure 8 shows a correlation between the mass rate of

diffusive combustion md and the injection quantity of gas

oil All Qd data shown in Figure 7 are plotted again in

Figure 8 It is clear that the mass rate md correlates well

with the injection quantity of gas oil, in other words, md is

strongly dependent upon injection quantity of gas oil and

its amount is about 70% of the injected gas oil

Figure 9 shows the relationship between the particulatesmatter in the exhaust gas and the parameter md/(1−XEGR)for accessing the effect of md and EGR on reduction in PM;where 1−XEGR of the denominator is adopted instead of

XEGR because the denominator is 0 at XEGR=0 It is clearfrom Figure 9 that the PM is strongly dominated by md/(1−

XEGR) If the EGR ratio is constant, the PM increases as themass rate of diffusive combustion md increases, and alsothe PM increases as the EGR ratio increases if md isconstant It cannot be simply determined whether the PMdecreases or increases because variation of md is dependentstrongly upon ignition delay In order to reduce the PM byincreasing the EGR ratio, the value of md/(1−XEGR) should

be decreased as shown in Figure 9 by two solid circles onthe correlation line This condition is written by the follow-ing equation;

i.e (2)where md0 is a value of the mass rate of diffusivecombustion in each condition without EGR (XEGR0).Figure 10 shows the ratio of diffusive combustion quan-

m d = Q d / Hu

m d / 1 ( – X EGR )< m d0 ( 1 – X EGR0 )

m d / m d0 <1 – X EGR

Figure 7 Change in cumulative heat release during

diffu-sive combustion Qd due to ignition delay Figure 8 Corelation between mass rate of diffusive com-bustion md and injection quantity of gas oil

Figure 9 Relationship between PM and md/(1−XEGR)

ENHANCEMENT OF NOx-PM TRADE-OFF IN A DIESEL ENGINE ADOPTING BIO-ETHANOL AND EGR 615

tity md/md0 due to the EGR ratio The right side of equation

(2) corresponds to the thick solid line with its gradient of

−1, in other words, if the value of md/md0 locates in the

region below the thick solid line, the PM can be reduced by

adopting EGR The experimental data shown in Figure 10

are identical to the ones shown in Figure 4 Only in the case

of 50% ethanol blend fuel EtOH50, PM could be reduced

by increasing the EGR ratio, and NOx, of course, could be

reduced simultaneously as shown Figure 11

3.4 Improved Results of PM-NOx Trade-off

Figure 11 shows the process of improvement in trade-off of

NOx-PM The ordinate denotes the specific particulates

matter PM and the abscissa denotes the specific emission

of NOx As shown in Figure 11, the original high value of

PM in the case of gas oil operation is reduced drastically by

blending ethanol about 50%, however, NOx increases a

little by the promotion of premixing In the case of 50%

ethanol blend fuel operation, by giving a small EGR ratio

of 0.2, both NOx and PM can be reduced simultaneously

Achieved very low levels of 1.5 g/kWh NOx and 0.02 g/

kWh PM fairly close to the post new-long term emission

standards in Japan could be achieved without deterioration

of the thermal efficiency by formation of the premixed

charge condition, resulting from the large ignition delay

4 CONCLUSIONSThe goal of the present study is to show one of theapproaches for achieving the combustion performance withlow NOx and PM combustion in a diesel engine byutilizing bio-ethanol and adopting EGR The gas oil wasblended with bio-ethanol for realizing a PCCI engine in thepresent experiment The concluding remarks obtained hereare as follows;

(1) Very low levels of NOx and PM could be achievedwithout deterioration of the thermal efficiency by pro-motion of premixing resulting from a large ignition delay.(2) PM is a function of the mass rate of diffusive com-bustion md and the EGR ratio, where md is dominated

by the injection quantity of gas oil

(3) Increase in the ethanol blend ratio results in a decrease

in the diffusive combustion quantity

(4) In the case of the 50% ethanol blend fuel, PM wasreduced by increasing the EGR ratio due to a remark-able increase in ignition delay

(5) Ignition delay increases with increase in the ethanolblend ratio, and also increases further by increasing theEGR ratio, resulting in promotion of premixing (6) A marked increase in ignition delay due to ethanolblending is based on lower cetane number and largerlatent heat of ethanol

ACKNOWLEDGEMENT− The authors wish to thank to Mr Nishjima, K., et al graduate students in Energy System Laboratory, Nagasaki University.

REFERENCESIshida, M., Tagai, T and Ueki, H (2004) Effects ofgasoline/gas-oil blend on smoke reduction in a dieselengine Reports of the Faculty of Engineering, Nagasaki University 34, 63, 9−15

Ishida, M., Jung, S., Ueki, H and Sakaguchi, D (2006).Effect of exhaust gas recirculation on combustion in aHCCI engine fuelled with DME/Natural gas Int Conf 2nd VAFSEP, Paper No 133, 61−66

Ishida, M., Jung, S H., Ueki, H and Sakaguchi, D (2007).Combustion characteristics of HCCI engines fuelledwith natural gas and DME 25th CIMAC, Paper No 171,

1−15

Ishida, M., Jung, S H., Ueki, H and Sakaguchi, D (2008).Experimental analysis of thermal efficiency improve-ment due to high EGR ratio in HCCI engines fuelledwith DME and natural gas COMODIA, 289−296.Jung, S., Ishida, M., Ueki, H and Onitsuka, S (2005).Ignition and combustion characteristics of DME pre-mixed with natural gas in a HCCI engine Internal Combustion Engine Symp 2005, Paper No 20056086 Jung, S H., Ishida, M., Ueki, H and Sakaguchi, D (2007)

Figure 10 Change in diffusive combustion quantity

normalized by initial value md/md0 due to EGR

Figure 11 Improvement in PM-NOx trade-off due to

ethanol blend and EGR

616 S JUNG et al.

Ignition characteristics of methanol and natural-gas in a

HCCI engine assisted by DME SAE Paper No

2007-01-1863, 1−6

Kamio, J., Kurotani, T., Kuzuoka, K., Kubo, Y., Taniguchi,

H and Hashimoto, K (2007) Study on HCCI-SI

com-bustion using fuels containing ethanol SAE Paper No

2007-01-4051

Li, T., Izumi, H., Shudo, T., Ogawa, H and Okabe, Y

(2007) Characterization of low temperature diesel

com-bustion with various dilution gases SAE Paper No

2007-01-0126

Li, T., Suzuki, M and Ogawa, H (2008) Effects of

in-cylinder temperature and fuel-air mixing on smokelesslow temperature diesel combustion COMODIA, Paper

No DE1-1, 135−142

Miyamoto, N., Chikahisa, T., Murayama, T and Sawyer, R.(1985) Description and analysis of diesel engine rate ofcombustion and performance using Wiebe’s functions

SAE Paper No 850107

Ogawa, H., Li, T., Miyamoto, N., Kido, S and Shimizu, H.(2006) Dependence of ultra-high EGR and low temper-ature diesel combustion on fuel injection conditions andcompression ratio SAE Paper No 2006-01-3386

International Journal of Automotive Technology , Vol 11, No 5, pp 617 − 623 (2010)

617

NUMERICAL STUDY OF A LIGHT-DUTY DIESEL ENGINE WITH A DUAL-LOOP EGR SYSTEM UNDER FREQUENT ENGINE OPERATING

CONDITIONS USING THE DOE METHOD

J PARK 1) , K S LEE 2)* , S SONG 3) and K M CHUN 3)

1)The Graduate School, Department of Mechanical Engineering, Yonsei University, Seoul 120-749, Korea

2)Department of Automotive Engineering, Kyonggi Institute of Technology, Gyeonggi 429-792, Korea

3)Department of Mechanical Engineering, Yonsei University, Seoul 120-749, Korea

(Received 25 February 2009; Revised 10 December 2009)

ABSTRACT− Exhaust gas recirculation (EGR) is an emission control technology that allows for a significant reduction in NOx emissions from light- and heavy-duty diesel engines The primary effects of EGR are a lower flame temperature and a lower oxygen concentration of the working fluid in the combustion chamber A high pressure loop (HPL) EGR is characterized by a fast response, especially at lower speeds, but is only applicable if the turbine upstream pressure is sufficiently higher than the boost pressure On the contrary, for the low pressure loop (LPL) EGR, a positive differential pressure between the turbine outlet and the compressor inlet is generally available However, a LPL EGR is characterized by

a slow response, especially at low and moderate speeds In this study, of the future types of EGR systems, the dual-loop EGR system (which has the combined features of the high-pressure loop EGR and the low-pressure loop EGR) was developed and was optimized under five selected operating conditions using a commercial engine simulation program (GT-POWER) and the DOE method Finally, significant improvements in the engine exhaust emissions and performance were obtained by controlling several major variables

KEY WORDS : Exhaust gas recirculation, Dual-loop EGR, Light-duty diesel engine, Design of experiments, 1D simulation, Boost pressure

1 INTRODUCTION

A diesel engine has many advantages in terms of its fuel

consumption, combustion efficiency and durability Also, a

diesel engine emits a relatively small amount of carbon

dioxide (CO2), carbon monoxide (CO) and hydrocarbons

(HC) However, diesel engines are a significant source of

NOx emissions and particulate matter emissions in urban

areas As the level of environmental concern increases, a

reduction of NOx emissions is one of the most important

tasks for the automotive industry to overcome

Addition-ally, future emission regulations will require a significant

reduction in both NOx emissions and particulate matter by

using EGR and aftertreatment systems A comparison

between experimental results for urban buses equipped

with urea SCR and EGR + DPF systems using diesel fuel

has been reported in previous studies (López et al., 2009)

Exhaust gas recirculation (EGR) is an emission control

technology that allows for a significant reduction in NOx

emissions from light- and heavy-duty diesel engines The

primary effects of EGR (Zheng et al., 2004) are a lower

flame temperature and a lower oxygen concentration of the

working fluid in the combustion chamber

Because a high pressure loop (HPL) EGR is zed by a fast response, especially at lower speeds, it is onlyapplicable if the turbine upstream pressure is sufficientlyhigher than the boost pressure For a low pressure loop(LPL) EGR, a positive differential pressure between theturbine outlet and the compressor inlet is generally requir-

characteri-ed However, the LPL EGR is characterized by a slowerresponse than that of HPL systems, especially at low andmoderate speeds

Due to the reinforced regulations, the exhaust gas circulation system is widely used and is believed to be aneffective method for NOx and PM reduction Regardingthe future types of EGR, a newly developed, dual-loop EGRsystem has become a common option to consider The experimental results of dual-loop EGR systems werereported in Cho et al. (2008), who studied a high-efficiencyclean combustion (HECC) engine compared to the HPL,LPL and dual-loop EGR at five operating conditions.For the dual-loop EGR system, it has been reported thatthe determination of the intake/exhaust air fraction (Wang,2008) and turbocharger matching (Mueller et al., 2005;Czarnowski et al., 2008) are important However, there arecomplex interactions between the variables that affect the

re-*Corresponding author. e-mail: leeks@kinst.ac.kr

618 J PARK, K S LEE, S SONG and K M CHUN

overall engine system Therefore, it is necessary to identify

the dominant variables for specific operating conditions

One of the available optimization methods, design of

ex-periments (DOE), can help identify the dominant variables

that have an effect on the dependent variables at specified

operating conditions Lee et al (2006) studied the low

pressure EGR optimization using the DOE method in a

heavy-duty diesel engine with respect to Euro 5 regulations

The dominant variables that had an effect on the torque,

NOx emissions and EGR rate were the EGR valve opening

rate, start of injection and injection mass In that study, the

optimized LPL EGR system achieved a 75% NOx reduction

with a 6% increase of the BSFC

In the present study, one of the future types of EGR (the

dual-loop EGR system) that has the combined features of a

high-pressure loop EGR and a low-pressure loop EGR was

developed and optimized to find the dominant parameters

for frequent engine operating conditions by using a

com-mercial engine simulation program (GT-POWER) and the

DOE method As a result, a validation was performed to

find the torque, EGR rate, NOx and BSFC and to

compen-sate for the torque and BSFC in the dual-loop EGR system

2 ENGINE MODEL

2.1 Engine Specification

The engine specification that was used for modeling is

summarized in Table 1 An original engine was equipped

with a variable geometry turbocharger (VGT), an

inter-cooler and a high-pressure loop EGR system The

operat-ing parameters included the engine operatoperat-ing speed, fuel

flow rate, ambient conditions, and combustion data

Addi-tionally, the length of the connecting rods, the distance

between the piston and pin, the compression ratio, and the

coefficient of friction were collected and entered into

GT-POWER Data sets of the valve diameters, valve timing,

injection timing and injection pressure were also acquired

These data are considered to be classified information from

the engine manufacturer and cannot be listed in detail Theengine operating conditions are summarized in Table 2.The five operating conditions that were used in theanalysis were chosen from frequently operated regions atemission test points that were provided by the enginemanufacturer

2.2 Engine Analysis Tool, GT-POWERSimulations were performed by using GT-POWER, which

is designed for steady-state and transient simulations andcan be used for analyses that involve engine and powertraincontrols The software is based on one-dimensional gasdynamics that represent the flow and heat transfer in thepiping system and in the other components of an enginesystem (Gamma Technologies, 2006) The complicatedshape of the intake and exhaust manifolds were convertedfrom 3D models (by originally using CATIA) into 1Dmodels using a 3D-discretization process The conversionprocess allows for faster and easier analysis of the gasflows and gas dynamics

The combustion model that was used was the injection diesel jet (DI-jet) model and it was primarily used

direct-to predict the burning rate and NOx emissions ously

simultane-2.3 Engine Model with a HPL EGRBased on experimental data, an engine model with an HPLEGR was designed The boost pressure was matched forthe appropriate turbocharger speeds based on the turbineand compressor maps The injection duration for a giveninjection timing, injection pressure, combustion pressureand temperature was determined Then, the back pressure

at the turbine downstream and the EGR valve openingwere determined The EGR rate, temperature and pressuredrop after the EGR cooler was monitored by using actuatorsand sensors Finally, the results of the simulation werecompared to the experimental data Figure 1 shows theengine model containing an HPL EGR

The percentage of exhaust gas recirculation (EGR(%)) isdefined as the percentage of the total intake mixture that isrecycled exhaust,

where [m i = m a + m f + m EGR], m EGR is the mass of EGR and

m a and m f are the mass of air and fuel, respectively

Table 1 Engine specifications

Cylinder arrangement 6cyl., V-type

Connecting rod length 159 mm

Wrist pin to crank offset 0.5 mm

Table 2 Engine operating conditions

NUMERICAL STUDY OF A LIGHT-DUTY DIESEL ENGINE WITH A DUAL-LOOP EGR SYSTEM 619

(Heywood, 1988)

2.4 Engine Model with a Dual-loop EGR System and

Optimization

Based on the HPL model, a dual-loop EGR model was

designed Compared to the HPL EGR system, the flap

valve opening rate was one of the most important variables

for the pressure difference at P2-P1 as shown in Figure 2

First, the dual-loop EGR simulation was performed

using a constant boost pressure The flap valve opening at

the tail pipe and the turbocharger RPM, which had effects

on the boost pressure and the back pressure using the

dual-loop EGR system, were selected as independent variables

In this case, the NOx reduction rate increased, but the

torque and BSFC decreased Then, an optimization was

performed to compensate for the torque loss and brake

specific fuel consumption (BSFC) by modifying the injection

mass, start of injection (SOI) and EGR valve opening rate

The results of the simulation were compared to the

out-comes of HPL and the dual-loop models with respect to the

torque, EGR rate, BSNOx, and BSFC

2.5 Design of Experiments, DOE

In this study, an optimization based on the DOE methodwas performed

There are primary variables that have significant effects

on the torque, BSNOx, BSFC, and EGR rate For thisstudy, six independent variables were selected: the HPLEGR valve opening diameter, LPL EGR valve openingFigure 1 Engine model containing a HPL EGR Figure 2 Engine model containing a dual-loop EGR

Table 4 Comparison between the experimental and simulation data in terms of the injection mass and maximum cylinderpressure

Case no ExperimentNormalized integrated injected mass (fraction)Simulation Maximum cylinder pressure (bar)Experiment Simulation

valve HPLLPL basebase 15% open 30% open15% open 30% open

SOI 3 CA adv 1.5 CA adv base.Flap valve 15% close Base 15% open

Full factorial 36 = 729

620 J PARK, K S LEE, S SONG and K M CHUN

diameter, injection mass, start of injection (SOI), flap valve

opening diameter at the tail pipe and turbocharger RPM

(TC RPM) The appropriate ranges were then set and the

DOE method was performed based on the full factorial

design

The torque, EGR rate, BSNOx, BSFC and boost

pre-ssure were selected as response variables The range of

each independent variable was selected based on the engine

design performance

Table 3 shows the control factors and levels of

optimi-zation for the dual EGR system

2.5.1 HPL & LPL EGR valves

If the EGR valve opens too much, it can cause torque loss

In this optimization, the maximum increase of the EGR

valve diameter was 30% (Lee, 2006) at the given operating

conditions from the values at a constant boost pressure

The given experimental data were selected at a base level,

which is shown as Level 1 in Table 3

2.5.2 Injection mass

A 5% increase in the injection mass was selected and an

increased injection mass had an effect on BSNOx

However, it typically degraded the BSFC The given

experimental data were selected at a base level, which is

shown as Level 1 in Table 3

2.5.3 SOI

In general, the injection starts faster than aTDC −25 ~−23

CA (Heywood, 1988) If the fuel is injected too early,

imperfect combustion could degrade the engine

perfor-mance In this optimization, the maximum advanced CA

that was selected was within 3o of the range of the current

value An advanced SOI could increase the torque without

any other changes in the variables

Also, the EGR rate could increase up to 10% The given

experimental data were selected at a base level, which is

shown as Level 1 in Table 3

2.5.4 Flap valve and TC RPM

For the dual-loop EGR system, the pressure difference at

P2-P1 in Figure 2 was affected by the interaction between

the flap valve and the TC RPM, which had a dominant

effect on the EGR rate and BSNOx for the dual-loop EGR

system

The TC RPM was chosen to maintain a boost pressure

that was based on the compressor map (Mueller et al.,

2005; Czarnowski et al., 2008) A positive sign in Table 3

indicates an increase in the rotational speed of the

turbo-charger shaft, whereas a negative sign indicates a decrease

in the rotational speed of the turbocharger shaft

3 RESULTS AND DISCUSSION

3.1 Experiment Vs Simulation for the HPL

Table 4 shows a comparison between the experimental data

and the simulation data in terms of the injection mass andmaximum cylinder pressure The total injection mass andrate of the main injection were given, but the rate of pilotinjections had to be determined by matching the injectionduration and pressure Figure 3 shows the torque, EGRrate, and NOx results of the simulation and the experiment The differences of each point were within ±5% and it wasshown that the simulation results are in good agreementwith the experimental results

3.2 Optimization of the Dual-loop EGR SystemBased on the DOE method, the response variables weredetermined at a constant boost pressure with a fixed HPLvalve diameter Then, the torque and BSFC compensationwas performed

3.2.1 Constant boost pressure with a fixed HPL valve meter

dia-The dual-loop EGR system optimization was performedbased on a constant boost pressure and a fixed HPL valveopening diameter to minimize torque loss

Table 5 shows the target boost pressures for the loop EGR system from the experimental data The inputvalue of the HPL valve opening diameter was the same asthat of the HPL model

dual-Figures 4 and 5 show the comparison between the

Figure 3 Comparison between the experimental data andthe simulation results (Torque, EGR rate and BSNOx)

Table 5 Target boost pressure

Target boost pressure 1.03 1.23 1.11 1.43 1.56

NUMERICAL STUDY OF A LIGHT-DUTY DIESEL ENGINE WITH A DUAL-LOOP EGR SYSTEM 621

cylinder pressure and heat release rate of the HPL and the

dual-loop system at a constant boost pressure in Case 5 An

increased EGR rate caused the cylinder peak pressure to

decrease in the dual-loop EGR system Also, a lower peak

heat release rate corresponded to lower NOx emissions

(Cho et al., 2008; Heywood, 1988)

Figure 6 shows the results from the dual-loop EGR

simulation at a constant boost pressure Compared to the

HPL model, an average decrease of 8% of the torque and

8% of the BSFC is observed

About 12% of the maximum torque loss (Case 3) and

about 11% of the maximum BSFC loss (Case 1) occurred

for the dual-loop EGR On the other hand, an average

reduction of about 60% of NOx was achieved

Additional-ly, a maximum reduction of 80% of NOx was achieved due

to the marked increase of the EGR rate (Case 2) It appears

that the mass of the LPL EGR portion had strong effects on

the total NOx reduction at a larger pressure difference

between the turbine downstream and compressor upstream

In Case 3, the NOx reduction rate was lower because of a

smaller pressure difference at P2-P1 in Figure 2

3.2.2 Optimization for constant torque and BSFC

To compensate for the torque and BSFC at a constant boostpressure condition, an optimization was performed tomaintain the original boost pressure (bar) By advancingthe SOI and increasing the injection mass, the torque andBSFC could be compensated for Table 6 shows the results

of the optimization for constant torque and BSFC withcontrolled variables

Figure 7 shows the simulation results of the HPL anddual-loop EGR at a constant boost pressure and theoptimized dual-loop EGR An average improvement of 8%

of the torque and 5% of the BSFC was achieved compared

to the dual-loop EGR system at a constant boost pressure.Furthermore, a higher NOx reduction efficiency appears ineach case, except for Case 4

In Case 4 (1556 RPM/BMEP 9.93 bar), it appears thatthe variables that affected the torque and the BSFC weredecoupled with the NOx reduction rate due to the relatively

Figure 4 Cylinder pressure of the HPL and dual-loop EGR

622 J PARK, K S LEE, S SONG and K M CHUN

high load conditions It was necessary to carefully control

the variables at high-load conditions

For the optimized dual-loop EGR system, a 60%

improvement in the deNOx efficiency was achieved by

increasing the EGR rates for all of the cases compared to

the results for the HPL system

4 CONCLUSIONS

In this study, an engine simulation was performed to optimize

the dual-loop EGR system at five different engine operating

conditions As a result, the dual-loop EGR system in a

light-duty diesel engine has the potential to satisfy future

emission regulations by controlling the dominant variables

for given operating conditions The details are as follows:

(1) An engine model for the HPL EGR was developed

based on the experimental data of five operating

condi-tions The calibrated simulations showed a ±5%

differ-ence from the experimental results

(2) For constant boost pressure conditions, an average

reduction of 60% of NOx was achieved in the

dual-loop EGR system compared to the results for the HPL

system However, approximately 8% of the torque loss

and 8% of the BSFC loss occurred

(3) To compensate for the torque and fuel consumption,

independent variables such as the start of injection and

the injection mass were selected as additional control

factors Comparing these variables to the dual-loop

EGR system at a constant boost pressure, an

improve-ment of approximately 8% of the torque and 5% of the

BSFC was achieved, except at high-load conditions

(Case 4)

(4) For the optimized dual-loop EGR system, a 60%improvement of deNOx efficiency was achieved byincreasing the EGR rate for all of the cases compared

to the results for the HPL system

ACKNOWLEDGEMENT− The authors gratefully acknowledge the financial support of the Ministry of Knowledge Economy of the Korean Government for this project.

REFERENCESCho, K., Han, M., Wagner, R M and Sluder, C S (2008).Mixed-source EGR for enabling high efficiency cleancombustion mode in a light-duty diesel engine SAE Paper No 2008010645

Czarnowski, R., Joergl, V., Weber, O., Shutty, J and Keller,

Table 6 Predicted values and validation results

NUMERICAL STUDY OF A LIGHT-DUTY DIESEL ENGINE WITH A DUAL-LOOP EGR SYSTEM 623

P (2008) Can future emissions limits be met with a

hybrid EGR system alone? 2008 Diesel

Engine-Effici-ency and Emissions Research (DEER) Conf

Gamma Technologies (2006) GT-POWER User’s Manual

Version 6.2

Heywood, J B (1988) Internal Combustion Engine

Fund-amentals International Edition McGraw-Hill New York

Singapore

Lee, S J., Lee, K S., Song, S H and Chun, K M (2006)

Low pressure loop EGR system analysis using simulation

and experimental investigation in heavy-duty diesel engine

Int J Automotive Technology 7, 6, 659−666

López, J M., Jiménez, F., Aparicio, F and Flores, N

(2009) On-road emissions from urban buses with SCR

+ Urea and EGR + DPF systems using diesel and diesel Transportation Research, Part D, 14, 1−5 Mueller, V., Christmann, R., Muenz, S and Gheorghiu, V.(2005) System structure and controller concept for anadvanced turbocharger/EGR system for a turbochargedpassenger car diesel engine SAE Paper No. 2005013888.Wang, J (2008) Air fraction estimation for multiple com-bustion mode diesel engines with dual-loop EGR systems.Control Engineering Practice, 16, 1479−1486

bio-Zheng, M., Reader, G T and Hawley, J G (2004) Dieselengine exhaust gas recirculation – A review on advancedand novel concepts Energy Conversion & Management,

45, 883−900

International Journal of Automotive Technology , Vol 11, No 5, pp 625 − 636 (2010)

625

OPTIMIZATION RESEARCH FOR A HIGH PRESSURE COMMON RAIL

DIESEL ENGINE BASED ON SIMULATION

Y LIU 1) , Y.-T ZHANG 2)* , T QIU 2) , X DING 2) and Q XIONG 2)

1)School of Mechanical and Electronic and Automobile Engineering, Beijing University of Civil Engineering and

Architecture, Beijing 100044, China

2)School of Mechanical and Vehicular Engineering, Beijing Institute of Technology, Beijing 100081, China

(Received 24 November 2008; Revised 20 September 2009)

ABSTRACT− A TP (Temperature Phase) model is presented to carry out optimization calculation for a high-pressure common rail diesel engine Temperature is the most important parameter in the TP model For the lower branch (when temperature T < 850 K) of the S-shaped curve (auto-ignition phase), a 6-step ad-hoc model with adjusted rate constants of n- heptane is used, referred to steady state assumption Steady state assumption is based on the observation that, due to very fast chemical processes in combustion problems, many chemical species and reactions are in a quasi-steady state or partial equilibrium When a species is assumed to be in the steady state, the corresponding differential equation can be replaced by

an algebraic relation, which reduces the computational costs For the middle branch (850 K ≤ T < 1100 K) of the S-shaped curve, a 4-step model is used to calculate the unstable phase For the upper branch (T ≥ 1100 K) of the S-shaped curve, one- step model is used because the one-step model has widely been used in descriptions of flame stability where it essentially serves as a model that produces a thin flame with strong temperature sensitivity When zone temperature T<1500 K, only the soot precursors –PAHs (Polycyclic aromatic hydrocarbons) is created and there is no soot emission When zone temperature

T ≥ 1500 K, PAHs and soot source terms (particle inception, surface growth, oxidation, coagulation) are calculated The TP model is then applied in multidimensional simulations to carry out optimizing, which reduces experiment cost The results of cylinder pressures, the corresponding heat release rates, NOx and soot with variation of injection time at constant rail pressure, variation of EGR-rate at constant rail pressure and variation of rail pressure at constant EGR-rate between simulation and experimental data are analyzed The results indicate that the TP model can carry out optimization and CFD (computational fluid dynamics) and can be a tool to calculate for a high-pressure common rail diesel engine.

KEY WORDS : High-pressure common rail diesel engine, Polycyclic aromatic hydrocarbons (PAHs), Temperature phase model, Optimization calculation

1 INTRODUCTION

At present, it is difficult in China to manufacture a

conv-entional distributed pump diesel engine without

after-treat-ment to EURO-II to EURO-III environafter-treat-mental standards

Matching the optimism combustion system for a

high-pressure common rail diesel engine is very important and

necessary work to improve fuel efficiency and emissions

The development of the combustion system, i.e

combus-tion chamber geometry, fuel injeccombus-tion system, and air

management system was in the past a rather pragmatic

practice largely based on experience Fang et al (2008)

used an optically accessible single-cylinder high speed

direct-injection (HSDI) diesel engine, equipped with a

Bosch common rail injection system, to study low

temper-ature Modulated Kinetics (MK) combustion with a

retard-ed single main injection They employretard-ed high-speretard-ed liquid

fuel Mie-scattering to investigate the liquid distribution and

evolution By carefully setting up the optics, they obtainedthree-dimensional images of fuel spray from both thebottom of the piston and the side window and measuredNOx emissions in the exhaust pipe They identified theinfluence of injection pressure, EGR rate, and injectiontiming on the combustion modes based on experiment.Kim et al. (2008) investigated the effects of injectionconditions and swirl on D.I diesel combustion using atransparent engine system They conducted a combustionanalysis and steady flow test to measure the heat releaserate due to cylinder pressure and the swirl ratio In addition,they obtained spray and diffusion flame images using ahigh speed camera and captured the LII and LIS imagesusing an ICCD camera to investigate the soot distribution

in the cylinder These experiments give lots of real data and

a convincing basis for engine design However, new logies related to fuel injection have dramatically increasedthe range over which the parameters affecting the perfor-mance of an engine can be varied, making such an ap-proach intractable when attempting to reach future stringent

techno-*Corresponding author. e-mail: youtong@bit.edu.cn

626 Y LIU et al.

emission targets at maintained performance levels

Three-dimensional simulations offer an insight into

phen-omena influencing performance and emissions behavior

that are very difficult or even impossible to study

experi-mentally The NOx mechanism is well understood by many

researchers However, despite more than a century of

studies, soot formation mechanism during combustion of

hydrocarbons remains a challenging problem for

combus-tion science For practical consideracombus-tions, a variety of

simplified soot models have therefore been proposed One

applied directly for engine simulation is a two-step model

proposed by Hiroyasu et al (Hiroyasu et al.,1983) Owing

to its ease of implementation into CFD codes, this model

and its modifications have acquired wide popularity in the

community engaged in multidimensional diesel combustion

simulations (Han et al.,1996) Nevertheless, Hiroyasu et

al.’s model is an empirical description, in which the particle

growth and dynamics of soot are not taken into account

Some others, following the idea of Gilyazetdinov’s work

on carbon black formation, postulated a monodisperse size

distribution of soot particles and then modeled the soot

formation by one or two equations: one for the particle

volume fraction and the other for the particle number

density Moss et al (1995) have demonstrated that this

approach can give encouraging results, but despite their

improvements, neither of the models has received much

attention because of their limited range of applicability A

common feature of the above soot models is that the soot

formation is linked explicitly to the fuel concentration,

which in many cases is not in agreement with experimental

evidence

Consequently, several authors have provided in-depth

insight into the physics and chemistry of NOx and soot

formation inside an engine, which is necessary if one

wishes to create a PCI (premixed compression ignition)

combustion strategy Kamimoto and Bae (Kamimoto and

Bae, 1988) first illustrated the soot and NOx forming

regions on an equivalence ratio/temperature plot for diesel

emission control This plot identifies the direction that one

must move to achieve PCI combustion: cooler combustion

temperatures Based on the theoretical insights, an

experi-mental investigation was conducted by Jacobs and Assanis

(Jacobs et al.,2007) to explore the theoretically proposed

possibility to simultaneously lower NOx and soot, as well

as run a smokeless rich operation, with PCI combustion

They found the resulting ultra-low concentrations of NOx

and soot are the result of low-temperature combustion, a

finding supported by estimated peak bulk gas temperatures

that are below 1500 K Additionally, the insensitivity of

soot formation on equivalence ratio when operating within

the low-temperature combustion regime has been

experi-mentally demonstrated via measured exhaust soot

concent-rations at rich conditions Therefore, for combustion

temper-atures lower than 1500 K, the potential exists for PCI

combustion without forming soot

Thus, there is a need to develop numerical calculation

based on accurate modeling and analysis based on the latestresearch results The present research focuses on obtaining

an understanding of the emission characteristics in a realhigh pressure common rail diesel engine through numericalsimulations First, the new shape of piston is showed Second,the NOx model and soot model are presented The TP(Temperature Phase) model for the autoignition is given inparticular detail Third, the experimental set-up and bound-ary conditions required for the calculations, as well as theoperating points investigated, will be discussed Next, theresults of the simulations and the measured data are com-pared In order to assess the performance of the matchingresults in terms of predicting cylinder pressure, heatrelease, oxides of nitrogen, and particulate emissions over awide range of part-load conditions, an extensive parameterstudy varying injection timing, EGR-rate and rail pressurehas been conducted

2 COMBUSTION CHAMBER STRUCTUREFigure 1 shows the shape of piston The biggest advantage

is that the new structure increases the compression ratiofrom 16 to 18.2.The simulation and experiment state thestructure can fit the common rail system and split injection

Figure 1 Combustion chamber structure

Figure 2 Three sides of the new piston

OPTIMIZATION RESEARCH FOR A HIGH PRESSURE COMMON RAIL DIESEL ENGINE BASED 627

and further increase the emission standard Figure 2 shows

the three sides of the new shape of piston

3 MODEL

3.1 NOx Model

In addition to correctly describing the ignition chemistry

and the high temperature kinetics it is the aim of the

reaction mechanism used in this investigation to capture

the processes of importance in forming pollutants such as

NOx and soot The NOx mechanism used in the current

study was developed by Bollig et al (1996) The mechanism

takes the following paths to NOx −formation and

con-sumption into account: (1) Thermal NO; (2) Prompt NO;

(3) Nitrous NO-formation (via N 2 O); (4) Oxidation of NO

to N 2 O; (5) NO-Reburn by hydrocarbon radicals and

amines (NHx) Generally, thermal NO (often referred to as

Zeldovich-NO) is by far the most significant contribution

to in-cylinder NOx in a diesel engine Reverse Zeldovich is

the most important mechanism in reducing NOx in the

postflame oxidation process However, if the temperature is

too low, the NOx-chemistry becomes subject to a state

denoted as frozen equilibrium This refers to a situation

where NOx exists in proportions far above those given by

the chemical equilibrium at the prevailing temperature The

reverse reactions are too slow to drive the concentration

towards the equilibrium Prompt NO (sometimes referred

to as Fennimore-NO) becomes important when the residence

time in hot regions gets shorter, as suggested by its name

The Reburn-NO mechanism has to be considered under

conditions of strong thermal NO-formation Hence, models

that only consider the thermal path tend to overpredict NO

under conditions where prompt NO plays a minor role

3.2 TP Model

3.2.1 Gas-phase chemistry

The gas-phase chemical mechanism used in the present

study is composed of three parts:

(1) the n-heptane auto-ignition kinetics;

(2) small hydrocarbon oxidation chemistry;

(3) the formation and oxidation of PAHs (up to two ringed

aromatics) and polyynes (to diacetylene)

Since auto-ignition is the first important event of

diffu-sion combustion in diesel conditions, we began to construct

a kinetic mechanism for n-heptane ignition based on the

detailed model proposed by Barths (Barths et al., 2001)

However, the size of Barths’s mechanism (involving 324

chemical species and 1650 reactions) is too large to use in

multidimensional turbulent diffusion computation Therefore,

only a few lumped or equivalent intermediate species (e.g.,

the first and second n-heptyl isomers were retained to

represent all the n-heptyl isomers) and some relevant key

reactions were selected on the basis of steady state

assump-tions These reactions are crucial for reproducing the

auto-ignition behavior of n-heptane/air mixtures from the low/

intermediate regime to the high-temperature regime Thehigh temperature oxidation of n-heptane proceeds from theattack of the fuel by H, OH and HO2 radicals to form n-heptyl radicals and through the break-up of these into C2H4,

CH3 and H radicals These are oxidized by reaction of theC1-C2-chemistry.The low temperature chemistry of loweraliphatic hydrocarbons is characterized by the degeneratedchain branching which may be illustrated by the followingsequence of reaction steps:

RH+O2 → R+HO2 (initiation)RH+OH → R+H2O (initiation)R+O2 ↔ RO2 (first O2-addition)

RO2 → R'O2H (internal H-abstraction)R'O2H+O2 ↔ O2R'O2H (second O2-addition)

O2R'O2H → HO2R''O2H (internal H-abstraction)For n-heptane, R is represented by C7H15, R'=C7H14 andR''=C7H13 This low temperature mechanism is no longervalid when the temperature increases beyond about 800K.The competition of the reverse reaction of the first andsecond O2-addition with the subsequent internal H-abstrac-tion reaction is the key to the understanding of the negativetemperature dependence of the ignition delay With theincreasing temperature, these reverse reactions become fasterthan their forward reactions, thereby stopping the reactionsequence A transition to the high temperature mechanismmust occur

The apparent negative temperature dependence of theintermediate branch shall be explained by discussing asimplified 4-step ad-hoc model with adjusted rate constants

of n-heptane ignition, which is written as

+H and I=HO2R''O+H2O P represents combustion of theproducts, with P=7CO2+8H2O

The first two reactions correspond to a two-step hightemperature scheme containing endothermic fuel decom-position into small hydrocarbons and the exothermic oxida-tion of these into the final combustion products The lasttwo steps represent the degenerated chain branching mech-anism discussed above Combining all steps up to theformulation of HO2R''O leads to the third global step of themodel The fourth step contains the chain branching andthe oxidation to the combustion products Only the thirdreaction is considered to be reversible The activation energy

of the backward reaction 3b is assumed much larger thanthat of the forward reaction 3f Therefore, at low temper-atures the backward reaction 3b is unimportant However,

628 Y LIU et al.

at the temperature around 850K and higher, the backward

reaction dominates over the forward reaction and thereby

decreases the relative importance of reactions 3f and 4 in

the mechanism This will explain the transition from the

low temperature to the high temperature branch A lot of

parameters can be found in our previous paper (Liu et al.,

2005) However, we only deduced and discussed these

equations in theory and did not use them in a real diesel

engine at that time Now we have used them in

three-dimension turbulent diffusion combustion with turbulence

movement in a diesel engine

These reaction steps, which constitute the first part of the

gas-phase kinetics, were integrated with the small

hydro-carbon chemistry model of Warnatz (Warnatz et al.,1996),

forming the main body of the gas phase mechanism for

n-heptane oxidation Some reactions extracted from the GRI

1.2 mechanism (Frenklach et al.,1995) were included in

the mechanism for completeness For the third part, the

reaction steps involved in the formation and oxidation of

PAHs and polyynes were taken from the mechanism of

Wang and Frenklach (1997) without modification

3.2.2 Soot inception/nucleation

In the present study, we have followed the ideas of Tao and

Golovitchev’s work, and assume that the soot particle

nu-cleation takes place at sites where diacetylene (C4H2) or

naphthalene (A2) exists The choice of these two species is

based on the argument that soot particle nucleation is not

merely linked to the pyrolysis of the hydrocarbon fuel, as

assumed by Lindstedt (Lindstedt and Louloudi, 2005), but

it also has its own, unique macro-molecular precursor

formation chemistry Strictly speaking, however, these two

reaction steps do not correctly represent the soot particle

nucleation process, because soot does not only contain

carbon atoms in its incipient particles, it also has

consider-able amounts of hydrogen In addition, there are

uncer-tainties regarding the determination of the rate constants of

these two reactions As will be shown below, the values of

the effective activation temperature for soot nucleation

used in Lindstedt’s model are too high for use in the present

study Unless a very high pre-exponential factor is

employ-ed, computations with Lindstedt’s nucleation model predict

a rather low soot mass concentration We have found that

appropriate values for the effective activation temperature

are below 5000 K, thus assigning a value of 1500 K for this

parameter in the present study because it provides

reason-able predictions compared with the availreason-able experimental

data, especially for the short time available for soot

for-mation immediately after ignition

3.2.3 Surface growth

We have adopted the idea of Frenklach and Wang’s active

site model (Wang and Frenklach et al., 1997) for the

present study However, the model is modified by the

inclusion of the irreversible reaction to account for the

activation of surface sites by OH radicals and the

deacti-vation of surface radical sites by H2O This reaction isfound to be critical in the soot formation in diesel spraycombustion and the rate constants are fully consistent withthose of the HACA reactions of PAHs Only part of the rateconstant for reaction is adjusted, to achieve agreement withexperimental data

3.2.4 Surface oxidationSoot oxidation is a heterogeneous process that takes place

on the surface of soot particles and depletes the carbonatoms accumulated in the particles However, since themechanism of this process is still poorly understood, it iscommonly described using empirical formulations such asthe Nagle and Strickland-Constable semi-empirical model(Nagle et al., 1962) to account for soot oxidation caused by

O2 attack This model is adopted in the present study Thesoot oxidation is also sensitive to other oxidants such as

OH radicals We include Neoh et al.’s OH oxidation model(Tao et al., 2004) and the rate, correlated to the collisionefficiency of OH radicals with the soot surface

3.2.5 Particle coagulationParticle coagulation causes the number of soot particles todecrease We assume a monodisperse size distribution ofthe particles, and then describe the coagulation by theSmoluchowski equation (Frenklach, 1985)

3.2.6 Soot source termsThe formation of soot can be subdivided into the process ofparticle inception, surface growth and oxidation, particlecoagulation As for the PAHs a statistical approach is used

to describe the size distribution function of the soot cles The moments of the size distribution of the size aredefined as:

where N i is the number density of particle i with a mass

m i=i·m 1, with m1 being the mass of the smallest unitoccurring in a soot particle The moment M 0 is equal to thetotal particle number density:

The moment M 1 can be related to the volume fraction,which defines the ratio of the volume occupied by sootparticles to the gas volume:

where ρ s=1800 kg/m3 is the density of soot

The source term is:

(4)with pb=particle inception, con=condensation, sg=surfacegrowth, ox=oxidation

OPTIMIZATION RESEARCH FOR A HIGH PRESSURE COMMON RAIL DIESEL ENGINE BASED 629

Particle inception can be modeled as the coagulation of

two PAH-molecules Coagulation of particles of the same

type can be described by Smoluchowski’s equation:

(5)This equation gives the change of the particle number in

size-class i as a function of time The first terms in

Equation (5) describes the formation of new particles from

smaller sized particles and the second terms the

consump-tion of particles in the i-th size-class by coagulation with

particles of all size classes The frequency factor β i,j in the

free molecular regime is given by:

where k B is the Boitzmann constant, µ i,j is the reduced mass,

r i is the radius of particles of class i and ε i,j is the size

dependent coagulation enhancement factor due to attractive

or repulsive forces between the particles The

Smolucho-wiski equation can be formulated for the particle inception

omitting the second terms:

Equation (7) can by multiplying with i r and a summation

over all size classes is described in terms of the moments of

the PAH size distribution (note the change of the upper

summation limit with respect to j):

One can rewrite (7) with the help of :

(9)with

where V iand m i are the volume and mass of particle of size

class i respectively, and ε i,j the coagulation enhancement

due to inter-particle forces The enhancement factor due to

van der Waals interaction is set to a constant value of 2.2

The mean soot particle diameter is given by

4 SIMULATIONS

4.1 Numerical Implementation

The soot model integrated with the complex chemistry

mechanism of gas-phase combustion is used to simulate

the soot formation in liquid spray combustion The 3V code (LIU and Zhang,2008) is used to solve the time-dependent conservation equations of total mass, momen-tum, energy, and species mass concentrations in the react-ing gas-phase mixture Spray dynamics is modeled usingthe discrete-particle technique, in which each computationalparticle represents a number of droplets of identical size,velocity, etc In addition to particle-turbulence interaction,spray atomization, droplet breakup, evaporation, collisionand coalescence are treated using spray sub-models for asingle-component, vaporizing fuel sprays are provided inthe original KIVA-3V code Turbulence modeling is based

KIVA-on the compressible k–ε model (accounting for dilatationeffect) modified to include the effects of fuel droplets Acharacteristic micro-mixing time scale is determined fromthe turbulence modeling and then further used for thechemistry-turbulence interaction All these sub-models im-plemented in the KIVA-3V code have been proven relevant

to the accurate simulations of diesel spray combustion

In the computations, the mass conservation and transportequation for the soot mass concentration is solved jointlywith the equations for the other gas phase species As thecombustion process of sprays considered in this work has ashort life span (less than 5 ms) taking place in the centralpart of the combustion chamber, we considered the therm-ophoretic effect is a less important factor to the sootformation To define accurately the total surface area ofsoot particles, as required for determination of the rates ofsurface reactions, an additional equation for the soot numberdensity is solved However, we postulate that this latterequation, related to the nucleation and coagulation pro-cesses, is affected only by the local conditions Thisassumption seems to be acceptable because many measure-ments have found surprisingly similar number densities indifferent flames (Dec et al.,1995), suggesting that the acro-scale convective transport and turbulent diffusion of thesoot number density can be neglected

4.2 Basis for the Numerical Simulation The computations started at IVC and ended at EVO Theinjection time unit from us to CAD (crank angle degree)can use the following form for a 4-stroke engine:

(12)where φis crank angle degree, n is speed (r/min), t is time(us)

The wall temperature (450K) was set such that lated and measured pressures matched during the compre-ssion phase before injection started The wall temperaturewas held constant during the computations The swirl wasset to 0.5 times the number of revolution of the engine (r/min) which corresponds to the measured swirl Theinjection nozzle was located on the axis of symmetry Sincethe shape of the piston is not on the axis of symmetry, 360o

calcu-was modeled which corresponds to x-offset 4mm and

630 Y LIU et al.

offset 1 mm Figure 3 shows the computational grid used in

the simulations containing 63,359 cells This corresponds

to a grid resolution of 0.9mm, 3.5mm and 30 in the radial,

axial and azimuthally directions respectively

5 EXPERIMENT SET-UP

Figure 4 shows a schematic of the test bench configuration

EGR was varied manually by operating a valve and the

boost was supplied from an external supercharge NOx was

measured with a Chemiluminescent Detector (CLD) from

Thermo Electron Corporation (Environmental Instruments

Model 10) with an error of less than 1% of the full scale

The response time was 1.5 sec and 1.7 sec for the NO- and

NOx -mode, respectively Soot was measured with an AVL

415 smokemeter In this measurement a well defined

volume is sampled from the exhaust stream and forced to

pass through a filter An opacity investigation of the filter

provides a measure of the amount of soot in the exhaust

Soot measurements of the kind described here essentially

only give an indication of the Dry Particulate Matter (DPM)

An induction air flow meter from Aerzen (max flow

capa-city 100 m 3/h) was used to measure the air mass flow A

compensation volume (100L) was included in the system

in order to reduce oscillations in the air mass flow The

crank angle encoder was a ROD 420, capable of resolving

a tenth of a crank angle degree The in-cylinder pressure

was measured using a flush mounted quartz sensor from

Kistler (type 6061B) The pressure transducer is very

accu-rate on a relative basis but does not directly yield absolute

values In this measurement, the reference was taken to be

atmospheric pressure and had to be corrected for the

applied boost pressure The peak motored cylinder pressure

is extremely sensitive to small changes in the pressure at

intake valve closing (IVC) Typically, the maximum error

in the pressure measured by the transducer is 0.1% of the

pressure limit of the transducer, which in this case is 250

bar This leads to an inaccuracy of 0.25bar in the pressure

at IVC For a typical operating point investigated here, this

corresponds to a 13bar error in peak motored pressure and

around 12 bar at a typical start of injection It is thereforeclear that any attempt at simulating the combustion has to

be preceded by a thorough analysis of the compressionstroke

In Table 1 some data on the engine is summarized Thetest engine is a four cylinder, 2.8 L 4JB1 DI engine Theinjection system is a second generation Bosch Common-Rail featuring a maximum injection pressure of 1600 bar.The injector is equipped with a six-hole nozzle, where eachhole has a diameter of 0.124 mm The injector is slightlyoffset (by 1.0 mm) to the center of the cylinder axis to allow

a better cooling of the narrow bridge between the exhaustvalves The ohmega-shaped bowl-in-piston is offset to thecylinder with x-offset 4 mm and y-offset 1 mm This madethe model 360-degrees, offering exact calculation of thefull geometry Whereas the engine experiments were per-formed using standard diesel fuel, a model fuel comprising

n-heptane was used to represent diesel in the simulation.Wang and Rutland (2007) showed that the model fueldisplays very similar performance and emission characteristics

to diesel

Table 2 summarizes the relevant 4JB1 engine operating

Figure 3 Computational grid used in the simulations

Figure 4 Schematic of the test bench configuration

Table 1 Engine specification

Spray cone angle 150 degreeInjector protrusion 1 mm

OPTIMIZATION RESEARCH FOR A HIGH PRESSURE COMMON RAIL DIESEL ENGINE BASED 631

data for the 13 load conditions of ESC Since the Euro III

standard (2000), the earlier steady-state engine test

ECER-49 has been replaced by two cycles: the European Stationary

Cycle (ESC) and the European Transient Cycle (ETC)

Smoke opacity is measured on the European Load Response

(ELR) test The following testing requirements apply for

Compression ignition diesel engines: (1) Euro III:

Conven-tional diesel engines: ESC/ELR test, Diesel engines with

“advanced aftertreatment” (NOx aftertreatment or DPFs)

and EEVs: ESC/ELR + ETC; (2) Euro IV and later: ESC/

ELR + ETC (http://www.dieselnet.com/standards/eu/hd.php)

ESC is the basic demand for Euro III and Euro IV but here

we discuss only one work condition due to word

limita-tions in this paper

Table 3 summarizes the relevant operating data for the

investigated load conditions (2000 rpm, 50% load, torque

135N·m and power 31.498 Kw) A parameter study,

varying the EGR-rate, injection timing and rail pressure

around a baseline case defined by the numbers in bold in

Table 2 (EGR-rate 30%, rail pressure 800 bar and Start OfInjection −2 after top dead center (ATDC)) will bepresented here While varying one parameter the otherswere held constant

An important input for the spray calculation is the tion rate profile To this end measurements were performedusing an Injection Rate Meter (EVI) This device, some-time referred to as the “Bosch-tube”, is operated such thatfuel is injected into a long (≈8m) tube filled with fuel Asthe fuel is injected, pressure fluctuations proportional to theinjection rate arise and thus a measure of the mass flow rate

injec-is obtained Measurements were performed at various tubepressures reflecting the variation in cylinder pressure duringinjection However, for the case investigated here, theinjection rate profile was rather insensitive to the backpressure The reason for this is that the injection velocityscales with the square root of the relative pressure differ-ence and the injection pressure is substantially higher thanthe back pressure Figure 5 shows the measured injectionrate profiles corresponding to rail pressures of 600 bar, 800bar and 1000 bar, respectively The back pressure was 50bar These shapes were used as input for the simulations to

be presented

6 RESULTS AND DISCUSSIONFor each of the simulations cylinder pressure, heat releaserate, NOx and soot are compared with measured values.The heat release rates were filtered, applying a so-calledButterworth-filter in order to remove high-frequent contri-butions to the signal Figure 6 illustrates the effect of filter-ing the raw data Whereas the computational results aregiven in units of gram soot per kg fuel, i.e Soot EmissionIndex (SEI), the experimental data was obtained in units ofFilter Smoke Number (FSN) Normally, smoke numbersare converted to soot concentrations in units of mg/m3 ap-plying some correlation (Muntean, 1999) Due to the un-certainties in these correlations at the low soot concent-rations considered here, such a conversion was not under-taken

Table 2 13 work conditions of ESC

Speed (r/min) Torque (N·m) Power(kW) Rail pressure(bar)

Start of Injections (ATDC) −10, −7, −4, −2, 2

Injection duration 6.0, 7.0, 8.0CA

Intake Valve Closing (IVC) 125 BTDC

Exhaust Valve opening (EVO) 126 ATDC

Rail pressure 600, 800, 1000bar

Figure 5 Injection rate profile

632 Y LIU et al.

6.1 Variation of Injection Time at Constant Rail Pressure

Figure 7 shows a comparison of measured and calculated

cylinder pressures and the corresponding heat release rate

for five different injection timings at a constant EGR-rate

of 30%, and rail pressure 800 bar As can be seen in Figure

6, the overall agreement in both ignition delay and peak

cylinder pressure is excellent The over-prediction of the

initial heat release discussed above is rather subtle, as seen

in Figure 7 It was found that more fuel had to be injected

at the most retarded injection timing, 2 degrees ATDC,

which was observed in the experiments as well However,

at this late injection, the simulations significantly predict the peak cylinder pressure

under-The results for NOx and soot are compiled in Figure 8

NOx is predicted very well indicating that its major sources

of influence are captured in the simulations However, thesoot prediction is not satisfactory especially when the starttime is -2ATDC

It is interesting to note that at the part load pointconsidered here, the cylinder temperatures in the expansionstroke are generally too low for the reverse Zeldovichmechanism to play a role reducing the levels of NOx This

is reflected in Figure 9 showing the NOx -history overdegree crank angle for the most advanced injection timing.Also included in the Figure 9 is the total NOx -productionterm, which apparently never goes below zero Regardingsoot, the overall trend observed in the experiments iscaptured qualitatively by the model Engine-out soot levelsare obtained as the difference between two large numbers,representing the processes of formation and oxidation,respectively This is one of the inherent problems makingquantitative soot predictions so challenging At injectiontimings before TDC, the familiar trend of increasing sootwith retarded injection timing is observed For the very lateinjection case, however, soot decreases again The explana-tion can be found by looking at the in-cylinder soot history,shown in Figure 9 Early injection timings are associatedwith substantial formation of soot followed by a ratherefficient oxidation For these cases engine-out soot is main-

ly controlled by the oxidation process When injecting afterTDC, the oxidation process is highly inefficient due to therapidly decreasing cylinder temperatures in the expansionstroke The simulations suggest that, in this case, onlyaround 50% of the soot formed gets subsequently oxidized.This should be compared to almost 90~96% for the mostFigure 6 Filter applied to heat release signal

Figure 7 Cylinder pressures and apparent heat release

OPTIMIZATION RESEARCH FOR A HIGH PRESSURE COMMON RAIL DIESEL ENGINE BASED 633

advanced injection timings The latter number is similar to

that observed by Schwarz et al (1999), who performed

optical diagnostics in a transparent heavy-duty DI diesel

engine However, despite the poor oxidation when

inject-ing late, the engine-out soot levels are relatively low This

is attributed to the small amounts of soot formed, as seen in

Figure 9 Thus, at late injection timings, engine-out soot

levels appear to be primarily controlled by the formation

process The fact that less soot is formed is in this case a

result of incomplete combustion, which is confirmed by the

experiments, showing a significant increase in unburned

hydrocarbons and carbon monoxide

6.2 Variation of EGR-rate at Constant Rail Pressure

Figure 10 shows the cylinder pressure traces and

corre-sponding heat release rates for the measured and simulated

parameter study varying the EGR-rate In all but the 50%

EGR-case, the heat release rate is very well predicted The

30% EGR-case was omitted here, since it was included in

the injection timing study

NOx is also very well predicted, as shown in Figure 11.The significant decrease when going from zero EGR to20% EGR is captured in the simulations Generally themodel slightly underpredicts NOx, except in the 50% EGR-case The latter is most likely due to the overestimatedinitial heat release for that case Measured soot increasescontinuously with increasing EGR-rate, as expected Thistrend is reproduced by the computations However, thesimulations fail to capture the significant difference inengine-out soot between EGR-rates of 40% and 50% Thereason for this may be that the oxidation of the bulk-gassoot, i.e soot formed in the center of the combustionchamber rather than at the walls, is overestimated In thecurrent model the main portion of the soot remaining at

Figure 10 Cylinder pressure traces and corresponding heat

634 Y LIU et al.

exhaust valve opening (EVO) originates from wall regions,

where flame quenching has occurred In general, more soot

is formed at higher EGR-rates, which is seen in Figure 11

and also in agreement with a previous analysis by Pitsch et

al (1996) This is due to the overall richer mixture, which

in the model translates to a displacement of the

stoichio-metric mixture fraction towards the lean The result is a

greater overlap between the flamelet profile representing

the first soot moment and the mixture fraction PDF This

behavior has also been observed by Hasse et al (2000)

6.3 Variation of Rail Pressure at Constant EGR-rate

Three different rail pressures were tested: 600, 800, and

1000 bar Figure 12 shows the pressure traces and heat

release rates for the different cases As one would expect,

the greatest heat release rate is observed for the

high-pressure case, although the difference to the 800 bar case is

rather subtle

Interestingly, there is no significant difference in ignition

delay between the three cases Higher injection pressures

produce finer sprays and overall leaner mixtures Hence,

when the mixture autoignites more mixture at

stoichio-metric is available in the high-pressure case The fact that

the mixture is more homogeneous at higher injection

pre-ssures has important implications for the soot, as displayed

in the lower diagram of Figure 13 The low soot emissions

at higher injection pressures are a result of less formation,

which is clearly demonstrated by Figure 13 The

computa-tions suggest a more modest decrease in soot with

increas-ing injection pressure than that observed in the

experi-ments Regarding the other pollutant considered here, it

may appear counter-intuitive that NOx, also, decreases with

increasing injection pressure, as seen in the upper diagram

of Figure 13 However, studying the heat release traces,

shown in Figure 12, it can be concluded that the

combus-tion duracombus-tion becomes shorter as the injeccombus-tion pressure

increases, i.e the residence time at high temperature isreduced Although the peak heat release rate is greater athigher injection pressures, which would normally lead tomore NOx, the reduction in combustion duration can not becompensated for This simultaneous reduction of NOx andsoot is one example of the ample potential of a common-rail injection system It should, however, be noted thatmore energy is required to operate the high-pressure fuelpump at higher injection pressures, which is reflected in theoverall fuel consumption According to the experiments,approximately 4% more fuel is required for the samepower output at a rail pressure of 1000 bar compared to the

NOx as a result of unburned fuel EGR appears to be themost adequate means of changing NOx and soot over awide range It is particularly interesting to note that thereare limits at which a small decrease in the exhaust concent-ration of one pollutant can only be achieved at the expense

of a significant increase of the other If an aftertreatmentstrategy is pursued to make the engine compliant withfuture emission legislation, one is faced with the question

as to whether to tune the engine so as to produce tory NOx -levels and reduce soot by means of a DieselParticulate Filter (DPF) or whether NOx should be reducedusing e.g a Lean NOx Catalyst at a soot level satisfying theemission targets Perhaps a combination of the two techni-ques will provide the solution The analysis performed heredemonstrates that new technologies, such as high pressure

satisfac-Figure 13 Comparison of measured and predicted NOx

OPTIMIZATION RESEARCH FOR A HIGH PRESSURE COMMON RAIL DIESEL ENGINE BASED 635

injection systems open up new horizons in optimizing the

combustion process The possibilities of reducing tail-pipe

emissions in the absence of any aftertreatment devices are

far from exhausted

7 CONCLUSIONS

(1) The Temperature Phase model for soot calculation was

applied to simulate a wide range of part load operating

conditions, varying injection timing, EGR-rate and

injec-tion pressure, in a four-cylinder version of the 4JB1

high pressure common-rail diesel engine, for which an

extensive set of experimental data exists Peak cylinder

pressure, apparent heat release rate, and emissions of

soot were very well reproduced by the model The

results indicate that the model can be used for

qualita-tive and even quantitaqualita-tive prediction of auto-ignition

and soot formation in diesel combustion

(2) The changed combustion chamber structure together

with the common rail system, flexible EGR and

injec-tion strategy can reduce raw emissions at the same time

as keeping original performance When engine speed is

2200 r/min for 50% load, the best optimistic result is

that rail pressure is 800 bar, EGR-Rate: 30% with SOI:

-2ATDC

(3) The three dimensional simulation model code capturing

the essential characteristics of the combustion process

can quickly deliver answers to questions regarding, for

example, the engine response to a certain fuel injection

strategy, bowl shape, or port geometry, thus providing

engine developers with an effective tool for cutting

lead times

ACKNOWLEDGEMENT− The authors thank Prof Norbert

Peters (RWTH-Aachen, Germany) for valuable discussions and

Prof Maggy McNorton (University of Glamorgan, UK) for

checking English grammar and spelling This project has been

supported by Beijing Natural Science Foundation (3102011),

Beijing Municipal Commission of Education Foundation

(KM200910016014) and Ministry of Housing and Urban-Rural

Development of the People’s Republic of China (MOHURD)

(2009-K8-5) and Funding Project for Academic Human

Re-sources Development in Institutions of Higher Learning Under

the Jurisdiction of Beijing Municipality (PHR(IHLB)) We would

like to extend our gratitude to those organizations that supported

this study.

REFERENCES

Barths, H (2001) Simulation of Diesel Engine and Gas

Turbine Combustion using Multiple Flamelets with

Detailed Chemistry Ph.D Dissertation RWTH-Aachen

Fang, T., Coverdill, R E., Lee, C.-F F and White, R A

(2008) Low temperature premixed combustion within a

small bore high speed direct injection (HSDI) optically

accessible diesel engine using a retarded single injection

Int J Automotive Technology 9, 5, 551−561

Feng, T., Valeri, I G and Jerzy, Chomiak (2004) A menological model for the prediction of soot formation

pheno-in diesel spray combustion Combustion and Flame,136,

Hasse, C., Bikas, G and Peters, N (2000) Modeling diesel combustion using the Eulerian particle flameletmodel SAE Paper No. 2000-01-2934

DI-Hewson, J C and Bollig, M (1996) Reduced mechanismsfor NOx emissions from hydrocarbon diffusion flames.26th Symp (Int.) Combustion, The Combustion Institute,Pittsburgh,2667−2672

Hiroyasu, H., Kadota, T and Arai, M (1983) Developmentand use of spray combustion modeling to predict dieselengine efficiency and pollutes emissions Bull JSME,

26, 569−575

Jacobs, T J and Assanis, D N (2007) The attainment ofpremixed compression ignition low-temperature com-bustion in a compression ignition direct injection engine.Proc Combustion Institute, 31, 2913−2920

Kamimoto, T and Bae, M (1988) High combustion ature for the reduction of particulate in diesel engines.SAE Paper No. 880423

temper-Kim, H M., Cho, W J and Lee, K H (2008) Effect ofinjection condition and swirl on D.I diesel combustion

in a transparent engine system Int J Automotive logy 9, 5, 535−541

Techno-Lindstedt, R P and Louloudi, S A (2005) Joint-scalartransported PDF modeling of soot formation and oxidation.Proc Combustion Institute, 30, 775−783

Liu, Y and Pei, P (2005) Analysis on ignition and tion of n-heptane in homogeneous systems Science in China, 48, 556−569

extinc-Liu, Y and Zhang, Y.-T (2008) Analysis on three sional computational fluid dynamics software based onKIVA-3V code Proc 3rd Int Conf Computer Science

dimen-& Education,Kaifeng, China, 259−263

Frenklach, M (1985) Dynamics of discrete distribution forSmoluchowski coagulation model J Colloid Interface Science, 108, 237–242

Moss, J B., Stewart, C D and Yong, K J (1995) ing soot formation and burnout in a high temperaturelaminar diffusion flame burning under oxygen-enrichedconditions. Combustion and Flame, 101, 491–500.Nagle, J and Strickland-Costable, R F (1962) Oxidation

Model-of carbon between 1000-2000-C, 5th Carbon Conf.,154M 164

Pitsch, H., Barths, H and Peters, N (1996) sional modeling of NOx and soot formation in DI dieselengines using detailed chemistry based on the interactive

Three-dimen-636 Y LIU et al.

flamelet approach SAE Paper No. 1996-01-2057

Schwarz, V., König, G., Dittrich, P and Binder, K (1999)

Analysis of mixture formation, combustion and pollutant

formation in HD diesel engines using modern optical

diagnostics and numerical simulation SAE Paper No.

1999-01-3647

Wang, H and Frenklach, M (1997) Detailed mechanism

and models of soot particle formation Combust and

Flame, 110, 173–221

Wang, Y and Rutland, C J (2007) Direct numerical lation of ignition in turbulent n-heptane liquid-fuel sprayjets Combustion and Flame,149, 353M 365

simu-Warnatz, J., Maas, U and Dibble, R W (1996) Combustion, Physical and Chemical Fundamentals, Modeling and Simulation, Experiments, Pollutant Formation SpringerVerlag

International Journal of Automotive Technology , Vol 11, No 5, pp 637 − 650 (2010)

637

MODIFIED ONE-STEP REACTION EQUATION FOR MODELING THE OXIDATION OF UNBURNED HYDROCARBONS IN ENGINE CONDITIONS

H KWON and K MIN *

School of Mechanical and Aerospace Engineering, Seoul National University, Seoul 151-742, Korea

(Received 7 September 2009; Revised 24 February 2010)

ABSTRACT− The oxidation of unburned hydrocarbons from piston crevices was modeled using a modified one-step reaction equation This new one-step oxidation model was developed by modifying the Arrhenius reaction rate coefficients of the conventional one-step reaction equation The predictions of the new one-step oxidation model agree well with the results of the detailed chemical reaction mechanism in terms of the 90% oxidation time of the fuel The effects of pressure and intermediate species in the burnt gas on the oxidation rate were also investigated and included as additional multiplying factors

in the modification of the equation To simulate the oxidation process of unburned hydrocarbons from a piston crevice, a dimensional computational mesh, based on the conventional engine geometry, was constructed with a fine mesh density at the regions of the piston crevice and cylinder wall The number of cell layers in the cylinder was controlled according to the piston motion to model the out-flow of unburned hydrocarbons from the piston crevice during the expansion stroke The effects of engine operational conditions on the oxidation rate were examined at several engine speeds and load conditions, and the sensitivity of the oxidation rate to the piston crevice volume was also evaluated Finally, the new one-step oxidation model was applied to a three-dimensional computational mesh that modeled the three-dimensional engine geometry and piston-valve motions to simulate the oxidation of unburned hydrocarbons in a real engine condition.

two-KEY WORDS : One-step reaction equation, Piston crevice, Unburned hydrocarbon, Oxidation, Practical engine condition

1 INTRODUCTION

To meet the upcoming, stringent standard for vehicle

tail-pipe hydrocarbon (HC) emissions, the reduction of

engine-out HC emissions is essential Among the several sources of

HC emissions, the combustion-chamber crevices appear to

be the largest contributors to engine-out HC emissions under

fully warmed engine conditions (Wentworth, 1971; Cheng et

al., 1993, Alkidas et al., 1995)

The major sources of unburned HC emissions and their

contribution to HC emissions in a warmed-up engine are as

follows (Cheng et al., 1993): crevices, about 40%; oil layers

and deposits, about 20% each; flame quenching and

in-cylinder liquid fuel effects, about 10% each; and the exhaust

valve leakage, less than 5% However, Wentworth reported

reductions in engine-out HC emissions as high as 74% when

piston crevices were virtually eliminated (Wentworth, 1971),

and Alkidas et al showed that combustion-chamber crevices

represent more than 50% of the HC emissions (Alkidas et

al., 1995) Irrespective of the percentage contribution, the

combustion-chamber crevices are recognized as the primary

sources of engine-out HC emissions

The combustion-chamber crevices include the

piston-ring-pack crevices, the head gasket crevices, the spark-plug

crevices and the valve-seat crevices The piston-ring-packcrevices, which are conveniently called the piston crevices,consist of the cylinder wall, piston-ring top-surface andpiston circumference Due to its largest volume portion andcomplex flow-reaction mechanism, many investigations ofthe combustion-chamber crevices are focused on the pistoncrevice (Min, 1994; Hellström and Chomik, 1995; Tonse,1996)

Crevices are narrow regions in the combustion chamber,into which the flame cannot penetrate During the compre-ssion stroke and the first stage of combustion, the cylinderpressure is increased, causing the in-cylinder gas to flowinto the crevices For a warmed-up engine, 4 to 6 percent ofthe total cylinder charge is trapped in the crevices at thepeak cylinder pressure (Min, 1994) The fuel trapped in thecrevices does not participate in the primary combustionprocess As the cylinder pressure decreases, and the pistonmoves down during the expansion process, most of thecrevice gas flows back into the combustion chamber, result-ing in high HC concentrations near the combustion chamberwall As the unburned HCs are mixed with hot burnt gases,some of them are oxidized during the expansion process orwithin the exhaust port, while the others escape the cylinderduring the exhaust process, which contribute to the engine-out HC emissions

Due to the difficulties in observing the oxidation process

*Corresponding author. e-mail: kdmin@snu.ac.kr

638 H KWON and K MIN

of piston crevice HCs by an experimental approach,

numeri-cal analysis is a useful tool to investigate this process

Computational fluid dynamics (CFD) has become

well-established over the last decades and is used widely in the

analysis of flow, combustion and emission prediction of

internal combustion engines In this study, the oxidation of

crevice HCs was simulated by three-dimensional CFD code

with a newly developed one-step oxidation model

Because the computing cost to perform an analysis

incor-porating a detailed chemical reaction model is too high, a

simplification of the model is essential In a previous study,

Hautman et al proposed a 4-step HC oxidation model,

involv-ing seven species of CnH2n+2, C2H4, CO, H2, O2, CO2 and H2O

(Hautman et al., 1981) However, Hautman’s 4-step model is

not appropriate for the modeling of unburned HC oxidation

because the reaction rate coefficients of the model were

developed under conditions different from those of internal

combustion engines Choi et al (2001) showed this

in-appropriateness in their study, and modified Hautman’s

coefficients for modeling the oxidation of propane (C3H8)

fuel Choi’s modified 4-step oxidation model was applied

successfully to the simulation of unburned HC oxidation in

three-dimensional computational meshes However, Choi’s

oxidation model was optimized for propane fuel, which is

not a gasoline surrogate, and the 4-step oxidation model still

took too much computational time for simulations in engine

geometry with a large number of computational cells

In this study, the simple one-step oxidation equation

pro-posed by Westbrook et al (Westbrook and Dryer, 1981)

was used for the modeling of the unburned HC oxidation

However, the original model of Westbrook et al is not

appropriate for the prediction of unburned HC oxidation in

practical engine conditions because it has been developed

in non-engine conditions and optimized for the prediction

of the flame speed and the flammability of the fuels The

reaction rate coefficients proposed by Westbrook et al

were modified to match the 90% burning time of the fuel

using the results of the detailed chemical reaction mechanism

of Curran(Curran et al., 2002) In addition, the effects of

increased pressure and intermediate species in the hot burnt

gas on the oxidation rate of unburned HCs were

investi-gated and included in the new one-step reaction equation

2 DEVELOPMENT OF A MODIFIED

ONE-STEP REACTION EQUATION

The modeling of unburned HC oxidation using a detailed

chemical reaction mechanism requires significant computing

time and effort The detailed chemical reaction mechanism

for HC fuels such as iso-octane, which has a large

mole-cular weight and complex molemole-cular structure, involves a

very large number of reaction steps and chemical species

In a numerical analysis of multi-dimensional reacting flow,

the number of governing equations for chemical species

must be sufficiently reduced to maintain a reasonable

ana-lysis run-time Therefore, the development of an effective

yet simplified oxidation model is required for the modeling

of unburned HC oxidation under practical engine conditions.The simplest form of a HC oxidation model is the one-step overall reaction model, which describes the globalrelation between reactants and products

(1)

In the above expression, the values of α, β, and χ arerelated by the number of carbon atoms in the fuel and theratio of the number of hydrogen to carbon atoms The rate

of conversion of fuel to CO2 and H2O depend on thetemperature and molecular concentrations of the reactants.Generally, the temperature dependency is described byArrhenius equation:

(2)where the pre-exponential factor A, the temperature ex-ponent β, and the activation energy E a (kcal/mol) are speci-fied Finally, the reaction rate is expressed as

(3)However, in many experimental measurements, the con-version rate of fuel to product species has been found toalso be dependent on the pressure condition, which rises toabove 30 bar in practical SI engines The conventionalsimple expression for the reaction rate in equation (3) doesnot consider this pressure dependency In addition, inter-mediate species exist in the hot bulk gases in the cylinder,which can affect the oxidation rate of unburned HCs.2.1 Modification of Reaction Rate CoefficientsWestbrook (Westbrook and Dryer, 1981) studied the sim-plified reaction mechanisms for the oxidation of HC fuelsand proposed two sets of reaction rate coefficients for theone-step overall oxidation of C8H18 fuel Table 1 showsthese two sets of reaction rate coefficients (Original A andB) However, the original coefficients of Westbrook et al.are not appropriate for the modeling of unburned HCoxidation because they were developed based on the experi-mental results obtained under conditions different fromthose of the engine cylinder and optimized mainly for theprediction of the flame speed and the flammability charac-teristics in a one-dimensional steady-state premixed flame.Thus, the original coefficients were adjusted to match theoxidation rate of iso-octane to that of the detailed chemicalmechanism under engine conditions The 90% fuel oxidationtimes were calculated under adiabatic and constant pre-ssure conditions using the CHEMKIN library (Kee et al.,1989) In these calculations, the fuel and oxidizer wereassumed to be well-mixed, and the system was modeled as

a non-dimensional WSR (Well-Stirred Reactor, Ando et

RT -

⎛ ⎞ Fuel [ ] a [ Oxidizer ] b

exp

=

MODIFIED ONE-STEP REACTION EQUATION FOR MODELING 639

mechanism was developed based on a series of

experi-ments at the initial pressure, which ranged from 1 to 45 atm,

and temperature, which ranged from 550 K to 1,700 K, and

validated through a jet-stirred reactor, flow reactors, shock

tubes and a motored engine These temperature and

pre-ssure conditions adequately covered the in-cylinder

temper-ature and pressure conditions during the expansion and

ex-haust processes of conventional SI engines The mechanism

consisted of 857 chemical species and 3,606 chemical

reac-tion steps

Figure 1 shows the 90% oxidation times of iso-octane,

calculated by the detailed reaction mechanism and three

one-step equations with the original and modified

coeffi-cients The initial pressure was 1 bar for the reference of

pressure effect, and the initial temperature was swept from

1,100 K to 2,000 K At temperatures below 1,100 K, the

time scale of 90% oxidation time was longer than that of

the engine expansion process For the time scales

associ-ated with the oxidation mechanism of unburned HCs,

Weiss and Keck (1981) applied a fast sampling valve in the

cylinder of a CFR engine and concluded that the complete

oxidation of the HCs occurred at gas temperatures above

1,250 K However, Medina (Medina et al., 1984) determined

that HCs exiting a small tube in the combustion-chamber

(simulated crevice) did not undergo rapid oxidation at

temperatures of up to 1,400 K using Shadowgraphy

photo-graphy and spontaneous Raman spectroscopy Additionally,

Eng (Eng et al., 1997) deduced from engine experiments thatthe critical cutoff temperature for post-flame HC consump-tion is near 1,500 K Additionally, a numerical simulationstudy of the post-flame oxidation of HCs by Wu (Wu andHochgreb, 1997) found the critical temperature to be around1,400 K Therefore, it was reasonable to set 1,100 K as thelower bound of oxidation temperature In Figure 1, theoriginal one-step models of Westbrook predicted the oxida-tion rate to be faster than that of Curran’s detailed mech-anism at temperatures lower than 1,300 K Above 1,300 K,the oxidation rate predicted by the original model was pre-dicted to be slower than the reference, and the differencesbetween the one-step and detailed models increased expon-entially Thus, the reaction rate coefficients of the originalone-step equation were modified to match the oxidationrate over the critical temperature range As can be seen inFigure 1, the resultant 90% oxidation time of the modifiedone-step equation shows good agreement with the result ofthe detailed mechanism over the wide temperature range.The values of the reaction rate coefficients for the originaland modified one-step reaction equations are listed in Table 1.2.2 Effect of Pressure

In general, the chemical reaction rate depends not only onthe temperature, but also the pressure conditions In thedetailed chemical mechanism of Curran et al., this pressuredependency was included by special fits or interpolations

in the mechanism Figure 2 shows the 90% oxidation times

of iso-octane, calculated by the detailed mechanism ofCurran (Curran et al., 2002) at pressure conditions rangingfrom 1 to 50 bar In a spark ignition engine, the peak value

of the in-cylinder pressure seldom rises above 50 bar, andthe pressure drops rapidly as the piston moves down afterthe primary flame propagation In Figure 2, the 90% oxida-tion time of iso-octane decreased as the initial temperaturewas elevated, and this tendency was maintained as the initialpressure was increased However, the slope of the oxida-tion time curve changed slightly as the initial pressure waselevated because the degree of acceleration in the reactionrate by pressure elevation was different at each temperatureconditions

To investigate the correlation between the oxidation rateand pressure effect, the oxidation times at elevated pressureconditions were calculated by the detailed chemical reactionmechanism and normalized by the oxidation time at 1 bar.Thus, the values in the y-axis of Figure 3 represent theratios of decrease in the 90% oxidation rate In Figure 3, wecan see that the ratios of decrease in the 90% oxidationtime showed an almost linear correlation with the pressureelevation The correlation lines were inclined upwardly atall temperature conditions, and the maximum slope of thelines was about 40 at the initial temperature of 1,100 K.This result means that the 90% oxidation time of iso-octanewas decreased by 1/40 as the pressure was elevated from

1 bar to 50 bar at that temperature, which implies a verylarge enhancement in the oxidation rate by the pressure

Table 1 Reaction rate coefficients of the original (Westbrook

and Dryer, 1981) and modified one-step reaction equations

for iso-octane (E a in kcal/mol)

Original A 4.6× 10 11 0.0 30.0 0.25 1.5

Original B 7.2× 10 12 0.0 40.0 0.25 1.5

Modified 4.6× 10 12 0.3 63.8 0.02 1.2

Figure 1 90% oxidation times of iso-octane at 1 bar by the

detailed mechanism and one-step equations with the original

and modified coefficients

640 H KWON and K MIN

effect However, the ratios of decrease in the oxidation time

were reduced as the initial temperature was elevated, and

when the initial temperature was higher than 1,500 K, the

oxidation rate was almost independent of the pressure and

the initial temperature condition

As the piston moves down during the expansion stroke,

unburned HCs in the piston crevice flow back into the

combustion-chamber and are mixed with the hot bulk gases

in the cylinder Choi and Park (Choi et al., 2001; Park and

Min, 2000) showed in their three-dimensional numeral

analysis that the temperature of the burnt gas near the

cylinder wall and piston surface ranges from 1,000 to 1,400 K

during the expansion stroke As can be seen in Figure 3,

this temperature range corresponds to the temperature

conditions at which the pressure effect was very large

Therefore, the model for the unburned HC oxidation must

consider the effect of raised pressure on the reaction rate In

this study, the results in Figure 3 were modeled by linear

equations at each temperature condition Then the

pre-exponential factor in the equation (2) (coefficient A) wasmultiplied by the value interpolated from the linear equations

at each temperature and pressure condition, so that theeffect of raised pressure was included into the one-stepoxidation model

2.3 Effect of Intermediate Species

In general, the oxidation products of a stoichiometric HCmixture are assumed to consist of CO2, H2O and inert gas,

N2 However, Haworth (Haworth et al., 2000) investigatedthe chemical composition of the burnt gases, and showedthat the burnt gas is composed of not only CO2, H2O andunburned fuel, but also some intermediate species, such

CO, H2, OH, O, and H Because these intermediate speciesare essential radicals and are known to supply chainbranching reactions in the oxidation of HC fuels (Sung andHuang, 2001), the effect of intermediate species existing inthe burnt gas must be considered in the modeling of post-flame chemical reactions, especially in the modeling ofunburned HC oxidation, where the unburned fuels areoxidized by being mixed with the bulk gases, which con-tain the intermediate species

To investigate the effect of intermediate species on theoxidation rate of iso-octane, the composition of intermediatespecies in the burnt gas must be determined first However,the chemical composition of the burnt gas cannot be deter-mined simply by the states after the primary combustion.During the expansion stroke, the temperature of burnt gas

in the cylinder falls continuously, and the chemical position of burnt gas is changed by recombination reactions

of intermediate species This change in the chemical position of burnt gas was calculated using the CHEMKINlibrary and the detailed chemical reaction mechanism, andthe results are shown in Figure 4 In the figure, the totalfraction of the intermediate species in the equilibrium mix-tures decreased exponentially as the equilibrium temper-ature decreased, and the portion of each intermediate speciesalso changed continuously At all equilibrium temperatures,

com-CO had the largest fraction among the intermediate species,and H2 had the second or third largest contribution de-pending on the temperature The contributions of radicals,such as H, O, were insignificant at temperatures below1,900 K

However, Heywood has produced the primary burnt ture charts for the working fluid in the combustion chamberduring the expansion process (Heywood, 1998) and conclud-

mix-ed that the recombination reactions of intermmix-ediate speciesare too slow to change the chemical composition of burntgas at temperatures below 1,700 K, i.e., the mixture com-position is frozen at 1,700 K Based on this result, thechemical composition of burnt gas was determined to be that

of the equilibrium mixture at 1,700 K, and the effect ofintermediate species in the mixture on the oxidation rate wasinvestigated by comparing the 90% oxidation times of twomodels of burnt gas composition: one including both ofintermediate species and complete oxidation products and

Figure 2 90% oxidation times of iso-octane at various

initial pressure conditions, calculated by detailed chemical

reaction mechanism

Figure 3 Decrease in 90% oxidation time according to the

pressure increase, normalized by the results at 1 bar

MODIFIED ONE-STEP REACTION EQUATION FOR MODELING 641

the other including only CO2 and H2O

Figure 5 shows the ratios of 90% oxidation times obtained

by using the two models of burnt gas composition with the

detailed chemical reaction mechanism of iso-octane The

volume fraction of the burnt gas in the total mixture, which

included both hot burnt gas and unburned crevice gas, was

varied from 0 to 99%, and the results were normalized by the

result at 0% burnt gas fraction Therefore, the y-value in

Figure 5 denotes the relative decreases in 90% oxidation

time, due to the inclusion of the intermediate species in the

oxidation process The initial temperature was varied from

1,100 K to 1,800 K, and the initial pressure was 1 bar In the

figure, at temperatures higher than 1,500 K, the intermediate

species in the burnt gas did not affect the 90% oxidation time

of the unburned HCs, which means that, the reaction rate

was fast enough to supply essential radical pools by itself

when the initial temperature was high, so that the

contri-bution of the intermediate species already existing in the

burnt gas was insignificant

However, as the initial temperature decreased, the effect

of the radical species in the burnt gas became considerable,

and the 90% oxidation time decreased in proportion to the

burnt gas fraction at temperatures lower than 1,500 K The

effect of the intermediate species reached its maximum at

1,200 K, and the 90% oxidation time was decreased by

more than 20% at that temperature when the burnt gas

fraction was increased up to 99% As the initial temperature

was decreased to lower than 1,200 K, the effect of the

intermediate species appeared to be saturated to a certain

level because temperature-controlled reaction steps were

delayed at low temperatures and dominated the overall

reaction time scale even though a sufficient amount of

radical species was supplied initially through mixing with

the burnt gas This trend appeared similarly at the elevated

pressure conditions

At all temperature conditions, the decrease in the

oxidation time became noticeable at burnt gas fractions

higher than 90 percent, which seemed to be relatively high.However, as discussed in a later section, this burnt gasfraction was easily met at the outer region of crevice gaslayers along the cylinder liner, where the unburned HCsdiffuse into the in-cylinder bulk gases Therefore, the effect

of the intermediate species in the burnt gas must be takeninto account in the modeling of oxidation of unburnedHCs In this study, the effect of the intermediate specieswas implemented in the pre-exponential factor of equation(2), which was multiplied by the value tabulated from theresults in Figure 5

3 2-D SIMULATION OF OXIDATION OF UNBURNED HYDROCARBONS

The oxidation of unburned HCs from the piston creviceduring the expansion stroke of an SI engine was simulatedusing a two-dimensional sector mesh The flow motions ofthe in-cylinder charge and crevice gas were solved by acommercial CFD code, and the oxidation rate of unburnedHCs was evaluated by the one-step reaction equation pro-posed in this study

3.1 Computational Setup3.1.1 Mesh configuration

A computational mesh with moving grids was constructed

to represent the piston motion in the combustion chamber

of a commercial SI engine, as specified in Table 2 Theengine cylinder can be assumed to be axisymmetric (Flowers

et al.,2003) and therefore, the mesh was constructed as a2-dimensional cross section with a center angle of 10o asshown in Figure 6 Both sections were defined as cyclicboundaries To observe the behavior of the unburnedhydrocarbons at the vicinity of the cylinder wall and thepiston crevice, computational meshes located near thewalls were refined, and the mesh size was graduallyincreased toward the center of the cylinder The dimensions

Figure 4 Chemical composition of burnt gas at various

equilibrium temperatures and pressure of 10 bar for a

stoichiometric mixture, calculated by the detailed chemical

reaction mechanism

Figure 5 Relative decrease in the 90% oxidation time ofiso-octane with burnt gas fraction in volume ratio at vari-ous initial temperatures, calculated by the detailed chemicalreaction mechanism

642 H KWON and K MIN

of the piston crevice were 0.35×6 mm in thickness and

depth, respectively, and its volume corresponded to 1.17%

of the clearance volume To maintain the aspect ratio of the

cylinder mesh during the compression and expansion stroke,

cell layers were deleted or added as the piston moved up or

down The maximum number of cells was 4,375 at BDC,

and the minimum was 1,200 at TDC

3.1.2 Combustion model and near-wall treatment

In the modeling of the oxidation of unburned HCs, it is

very important to predict accurately the in-cylinder

condi-tions after the primary combustion, especially the

temper-ature distributions at the regions near the walls, because the

oxidation rate of HCs is primarily dominated by the

temper-ature conditions In previous studies (Min, 1994; Wu and

Gochgreb, 1997; Choi et al.,2001) on the numerical

ana-lysis of the post-flame oxidation of unburned HCs, the

temperature field after the primary combustion was based

on several assumptions and simplifications, which may

result in an over- or under-estimation of the oxidation rate

of the unburned HCs

In this study, a three-dimensional premixed turbulent

combustion model, as proposed by Weller (1993) and

modi-fied by Kwon (Kwon et al., 2005), was used to simulate the

premixed turbulent combustion of the engine This bustion model has been well validated over a range ofengine speeds, loads, equivalence ratios and spark timing(Heel et al., 1998) In this model, premixed flame propa-gation is described by solving the transport equation of theprogress variable , which is equal to 0 in the fully un-burned gas and to 1 in the fully burnt gas Equation (4)shows the transport equation of progress variable and itssource terms, which represents fuel consumption, ignition,and wall quenching respectively Detailed descriptions can

com-be found in a previous paper (Kwon et al.,2005)

(4)Wall heat transfer is another important factor for theaccurate prediction of the temperature field in an SI engine

In this study, the modified wall function proposed byKleemann (Kleemann et al., 2001) was used to simulatethe heat transfer between the in-cylinder gas and thecombustion chamber surfaces This model is capable ofcapturing the effect of high temperature on the wall heattransfer, which increases the driving force for the heattransfer and further reduces the thermal resistance in theboundary layer via changes to the distributions of thethermo physical properties within the boundary layer InKleemann’s study, the dimensionless wall distance y +,velocity u +, and temperature T + of the conventional wallfunction were transformed to the new dimensionless vari-ables η +, ψ +, and Θ + respectively to reflect the density andviscosity variation effects across the boundary layer, andthe model showed substantially improved correlation withthe measured heat fluxes

3.1.3 Calculation conditions

To solve the turbulent flow field in the cylinder, STAR-CD,

a commercial CFD code for multipurpose thermo fluidanalysis, was used Several sub-models were incorporatedinto STAR-CD for the modeling of turbulent combustion,such as modified wall heat transfer and unburned HCoxidation For the turbulent model, the standard k-ε modelwas used, and thermodynamic and transport properties,such as the viscosity, thermal conductivity, specific heat,diffusion coefficients, and Schmidt numbers, were obtainedfrom the CHEMKIN thermodynamic and transport pro-perty database

The calculations began at −134o CA, which corresponded

to the intake valve closing, and ended at 146o CA, whichcorresponded to the exhaust valve opening, to simulate thecompression, flame propagation, and expansion processes

of the engine The initial temperature and pressure conditions

at IVC were obtained from the one-dimensional cyclesimulations, using the GT-POWER code, which reproducedthe gas exchange and combustion process of the engine(Lee et al., 2007) The base operational conditions, with thewarmed-up wall temperatures of the computational engine,are listed in Table 3

c˜

∂ ρc˜ ( )

∂t - ∇ ρuc˜ + ⋅ ( ) ∇ ρD˜ – ⋅ ( c ∇c˜ ) = Sc˜+ + S ig S Q

Table 2 Specifications of the modeled engine

MODIFIED ONE-STEP REACTION EQUATION FOR MODELING 643

3.2 Results and Discussion

3.2.1 Oxidation of the unburned HCs at the base condition

At the base operation condition with warmed-up wall

temperatures, the oxidation of unburned HCs, from piston

crevice, was simulated

In Figure 7, the velocities of the piston and crevice gas,

and the velocity differences between the two are shown to

investigate the out-flow motion of the crevice gas From

the figure, we can see that the velocity of the

downward-moving piston was faster than that of the crevice flowing gas, and thus the piston crevice gas was stretchedinto a thin layer along the cylinder liner Figure 8 shows thevelocity field at the piston crevice region at ATDC 70o CA.The left is the vector sum of x-y-z components, and theright is the magnitude of the velocity vector From thefigure, the flow motion of the crevice gas was dominated

out-by the velocity component in the z-direction, which wasderived by the downward movement of the piston Themagnitude of velocity decreased drastically at the vicinity

of the cylinder liner because of the non-slip condition at thewall, which promoted the generation of a thin layer ofunburned HCs on the cylinder liner

The primary mechanism of oxidation of unburned HCs

is the heat and species exchange between hot in-cylinderbulk gases and thin crevice gas layers The unburned fuelalong the cylinder liner diffuses into the hot burnt gasregion by the concentration gradient, and the heat from thehot cylinder charge diffuses into the piston crevice gas bythe temperature gradient

Figure 9 shows the fuel mass fraction and temperaturedistribution at the vicinity of the piston crevice and thecylinder wall At the beginning of the expansion process,the temperature of the in-cylinder charge was very high andenough to oxidize the unburned fuel as soon as it flowedout from the piston crevice Thus, the length of theunburned fuel layer was short, and the diffusion of theunburned fuel layer into the cylinder was not significant, asshown in the first plot of fuel mass fraction in Figure 9 at

Table 3 Base operation condition of computational engine

Intake pressure (bar) 0.4

Exhaust pressure (bar) 1.0

Wall temperature (K) Head/Liner373 Piston450

Figure 7 Difference in the velocity of piston and crevice

gas

Figure 8 Velocity field at the vicinity of the piston crevice

in a vector and contour plot at ATDC 70o CA, 1,500 rpm

Figure 9 Fuel mass fraction and temperature field at thevicinity of piston crevice at ATDC 40, 60, 80, 100 and 120o

CA, 1,500 rpm

644 H KWON and K MIN

40° CA ATDC The heat released from the oxidation raised

the temperature locally, and distorted the thermal boundary

layer on the cylinder wall at the vicinity of the crevice In

the two plots for the temperature field at ATDC 40o and 60o

CA, hot spots were found located at the outer region of the

thin layer This local temperature rise promoted the

oxidation of following crevice gases

As the piston kept moving down, the burnt gas

temper-ature fell, and the oxidation rate of the crevice gas slowed

down At this time, the rate of heat release from the

unburned fuel oxidation was not fast enough to maintain

the hot spot against the heat transfer to the surroundings,

and the shape of the temperature boundary quickly returned

to normal The temperature field at ATDC 80o CA, as

shown in Figure 9, was stable even though the unburned

HCs were being oxidized continuously After this, a large

portion of unburned fuel from the piston crevice survived

the oxidation process, forming an unburned fuel layer,

which was much thicker and longer than that at the early

stage of the expansion stroke

At the later stage of the expansion stroke, the

ature of the bulk gas fell below the critical cutoff

temper-ature In this condition, most of the unburned fuel did not

undergo the oxidation process, and formed a very thick and

long diffusion layer At 120o CA ATDC, the diffusion

thickness of the unburned fuel was larger than the

dimen-sion of the piston crevice by about three times, as shown in

Figure 9

In the early stage of the unburned HC oxidation process

when the oxidation rate was very fast, the diffusion of

unburned fuel was very weak and appeared to be

concent-rated in the narrow region located right above the piston

crevice However, as the piston moved down continuously,

the diffusion layer broadened, and the diffusion occurred

over the long surface of the cylinder wall In Figure 9, the

contour plot of the fuel mass fraction at 120o CA ATDC

shows that the mass fraction of unburned decreased

gradu-ally from a stoichiometric value of 0.062 to zero, with a

distribution across the mixing zones, and the burnt gas

fractions became larger than 90% along the outer region of

the diffusion layer As mentioned above, the effect of

inter-mediate species on the oxidation rate became noticeable as

the burnt gas fraction exceeded 90% In addition, the

temperature at the diffusive mixing zone fitted into the

conditions given in Figure 5, where the effect of

inter-mediate species was maximized Therefore, we expect the

modified one-step equation in this study to be able predict

the oxidation rate of unburned HCs more accurately by

considering the effect of intermediate species

The mass fraction curve of iso-octane in the cylinder

during the expansion process is shown at Figure 10 The

sub-model for HC oxidation was activated at 40o CA

ATCD, when the flame front arrived at the end of the

combustion chamber, and the mean value of the progress

variable in cylinder was about 0.97, which implied that the

combustion chamber was completely filled by fully burnt

gas, except at the piston crevice region From the figure,the mass fractions from the two one-step equations ofWestbrook were almost identical, and the oxidation rateswere predicted to be much smaller than that of the modifiedone-step model As shown in Figure 1, the 90% oxidationtimes of the original one-step models were much longerthan that of the modified one-step model, so that thepredicted oxidation rates had large differences even withthe same in-cylinder conditions The resulting oxidationrates of Westbrook’s original models 1 and 2 were 15.0%and 13.4%, respectively, while the modified one-stepmodel predicted an oxidation rate of 61.5% This resultmeans that 61.5% of the total mass of unburned HCs wasoxidized before the exhaust valves were opened Thisresult agrees fairly well with the result of Hamrin et al.

(Hamrin and Heywood, 1995), where 68% of the HCsoriginated in the combustion chamber crevices underwentin-cylinder oxidation Generally, there exists a transitionpoint, where the oxidation is no longer effective, and much

of the crevice HCs survive the oxidation In the result of themodified one-step equation, this transition point can befound at around 90° CA ATCD, which is in qualitativeagreement with the previous experimental findings ofWeiss(Weiss and Keck, 1981), and Min(Min and Cheng,1995) There were no significant transitions in the twomass fraction profiles obtained from Westbrook’s originalone-step equations

3.2.2 Effect of the engine operational conditionsDuring the cold start of an engine, the cylinder wall is atambient temperature, which deteriorates the oxidation ofunburned HCs Additionally, heat transfer to the wall isincreased, causing lower burnt gas temperature Thus, weinvestigated the effect of the wall temperature at the baseoperation condition The temperature of the piston topsurface and the cylinder wall for the cold-start conditionwere set at the levels given in Table 4 The initial temper-

Figure 10 Mass fraction of iso-octane during the expansionprocess calculated by the three one-step oxidation models

at 1,500 rpm and 0.4 bar

MODIFIED ONE-STEP REACTION EQUATION FOR MODELING 645

ature of the in-cylinder charge at the start of compression

with cold walls was obtained by a 1-D cycle simulation At

the cold start condition, the oxidation rate of unburned HCs

was found to be 44.7%, which was lower than that of the

warmed-up condition by about 17%

To investigate the effect of engine load on the oxidation

rate of the crevice HC, three engine load conditions of 0.4,

0.6 and 0.8 bar IMAP (intake manifold absolute pressure)

were considered An engine operating at high load

condi-tion emits large amounts of unburned HCs when the

ex-haust valve opens because the initial mass of unburned

HCs in the piston crevice increases due to the increased

density at elevated cylinder pressure In this study, the peak

pressure in the cylinder was computed to increase from 17

to 28 bar as IMAP increased from 0.4 to 0.8 bar In Table 5,

the initial mass of unburned HCs at the 0.8 bar IMAP

condition was 2.56 times that at the 0.4 bar IMAP

condi-tion However, at the time of the exhaust valve opening, the

remaining mass of unburned HCs at the 0.8 bar condition

was 1.62 times that at the 0.4 bar condition, which was

decreased from the initial ratio of 2.56 This lower ratio

was due to the oxidation of unburned HCs being enhanced

at elevated pressure conditions In addition, the greater the

HC outflow from the crevice, the greater the amount of

heat released, which resulted in the local temperature rising

with increased oxidation of the crevice gases The resultant

oxidation rates were 72.1 and 77.3% at the 0.6 and 0.8 bar

IMAP conditions, respectively

The effect of the engine speed on the oxidation rate wasexamined at 1,500, 2,500 and 3,500 rpm conditions En-gine speed did not affect the in-cylinder temperature orpressure as much as the engine load condition did Neverthe-less, the burnt gas temperature and the in-cylinder pressureincreased a little as the engine speed increased, as shown inTable 6 As mentioned above, increased temperature andpressure can promote the oxidation of the unburned HCs.However, the important factor in engine speed variations isthat the time available for heat transfer and oxidationreaction for each cycle decreases proportionally as theengine speed increases Another important factor is thevelocity difference between the piston and crevice gas Asshown in Figure 7, crevice gases were driven out of thepiston crevice by the velocity difference and non-slipconditions at the walls As the engine speed increased, thepiston moved down faster, causing a great amount ofcrevice gas to flow out from the crevice, which enhancedthe oxidation rate Therefore, there existed a competitionbetween the two factors, leading to decreased residencetime and increased flow rates of the crevice gas As aresult, the oxidation rate of unburned HCs was 63.4% at2,500 rpm, which was slightly increased, and 60.1% at3,500 rpm, which was slightly decreased from the basecondition

3.2.3 Effect of crevice dimension variationsThe sensitivity of oxidation rate to the crevice dimensionwas investigated by doubling and halving the crevicevolume at the base operational condition To achieve thesechanges in volume, the length of piston crevice waschanged to 12 mm and 3 mm, while the thickness of thecrevice in the cylinder radial direction was maintained atthe original dimension

Figure 11 shows the mass of the unburned HCs flowingout of the piston crevice at three crevice volume conditions.Although the velocity of the piston was the same at allcrevice dimensions, greater amounts of crevice gas flowed

Table 4 Effect of engine wall temperatures on the oxidation

rate at 1500 rpm and 0.4 bar IMAP

Cylinder wall (K) top (K)Piston Oxidationrate (%)

646 H KWON and K MIN

out of the piston crevice at the longer crevice size because

of the non-slip condition at the cylinder walls The mass

flow rate of unburned HCs was found to be linearly

proportional to the crevice dimension, which implied that

larger crevice dimensions promoted the local temperature

rise due to the oxidation of HCs In Table 7, the amount of

unburned HCs remaining at EVO increased as the crevice

size was increased, but the oxidation rate was also

en-hanced at larger crevice dimensions At halved crevice

dimensions, the relative amount of unburned HCs at EVO

was 0.82, while the initial ratio was 0.5, which implied that

the oxidation of unburned HCs deteriorated In contrast, at

doubled crevice dimensions, the ratio of the amount of HCs

at EVO was 1.45, which was decreased from the initial

ratio of 2.0 The resultant oxidation rate was 43.3% at

halved crevice dimensions and 74.7% at doubled crevice

dimensions

The sensitivity of amount of unburned HCs at EVO to

the piston crevice volume was evaluated to be 0.44, as

shown in Figure 12 In previous studies, the sensitivity of

HC emissions to the piston crevice volume was 0.2 for Min

(Min et al., 1994), 0.47 for Alkidas (Alkidas et al., 1995),

and 0.6 for an earlier study of Wentworth (1971) The

sensitivity varied widely in the literatures, and this is

believed to be a result of differences in the fuel type, the

engine conditions, and the geometry of the engine

com-bustion chamber, which affected the oxidation rate of

un-burned HCs However, the linearity of the sensitivity wasfound in all of the studies, which agrees with the result ofthe modified one-step model proposed in this study

4 SIMULATION OF OXIDATION OF UNBURNED HYDROCARBON In 3-D ENGINE GEOMETRY4.1 Computational Setup

To investigate the oxidation of unburned HCs in a realengine geometry, a three-dimensional computational meshwas constructed based on a conventional SI engine, whichwas first modeled in the two-dimensional simulations de-scribed above The computational domain of the meshincluded not only the combustion chamber, but also theexhaust ports to consider the in-port oxidation of unburnedHCs during the exhaust stroke The dimensions of thepiston crevice were not changed from those of the previous2-D simulation, which were defined as 0.35×6 mm in theradial and longitudinal directions, with 5×20 cell layers foreach direction For the wall to function effectively, thecylinder wall was wrapped by 3 cell layers with a totalthickness of 0.5 mm to maintain the value of y +, the non-dimensional wall distance, in the appropriate range Figure

13 shows the shape of the overall computational domainand mesh distributions near the walls The motions of thepiston and intake/exhaust valves were considered by mov-ing the cell vertices in accordance with the crank rotation,and the number of cell layers located in the cylinder aboveand below the valves was controlled to maintain the aspectratio of the cells The total number of cells was 576,000 atBDC, and 359,000 at TDC, which included 155,000 and129,000 cells for the intake and the exhaust ports, respec-tively

Other aspects, such as the engine’s operational tion, turbulent flow, and a premixed combustion model,were the same as those of previous 2-D simulations

condi-Table 7 Effect of crevice volume on the oxidation rate at

1,500 rpm and 0.4 bar IMAP

Figure 12 Sensitivity of the amount of unburned HC

remaining at EVO to piston crevice volume

Figure 13 Three-dimensional engine mesh and mesh stribution near the walls

di-MODIFIED ONE-STEP REACTION EQUATION FOR MODELING 647

4.2 Results and Discussion

As mentioned above, the state of the in-cylinder charge

must be predicted accurately in the first place to achieve

accurate modeling of the oxidation of unburned HCs

Figure 14 shows the cylinder pressure curves obtained from

engine measurements and three-dimensional CFD

simula-tions The calculation began at 34o CA BBDC when the

exhaust valves were opened and continued for two engine

cycles to exclude the effect of uncertainties in the initial

states of charges in the cylinder and ports at the start of

calculation Mass flow rates at the intake and exhaust

boundaries were obtained from a 1-D cycle simulation (Lee

et al., 2007) From the figure, the pressure curve predicted

by our 3-D simulation is in good agreement with the one

measured in the target engine, which confirms that the gas

exchange and combustion process of the engine was

successfully modeled in this study

Based on this prediction, the oxidation of unburned HCs

during the expansion and exhaust strokes of the engine was

simulated Figure 15 shows the behavior of unburned HCs

in terms of the mass fraction of the fuel in the interval of 60o

CA after firingTDC As discussed in the 2-D simulations,unburned HCs flowed out of the piston crevices as thepiston moved down, forming a thin layer, and then diffusedinto the combustion chamber At the early stages of theexpansion stroke, the temperature of the in-cylinder burntgas was sufficiently high to oxidize the unburned HCs assoon as they escaped from the piston crevices and con-tacted with the hot bulk gases Heat released from this earlyoxidation caused the temperature to rise locally at thevicinity of the piston crevices, and also promoted theoxidation of the accompanying crevice gases Therefore,the growth of the diffusion layer was very weak at thisstage From Figure 15, we can see that the distribution ofunburned HC was concentrated at the piston crevice region

at 10o CA ATCD

As the piston moved down continuously, the oxidationrate of unburned HCs became slow, and long and thicklayers of unburned HCs began to develop along the cylinderwall surface, as shown in the second plot of Figure 15.When the piston arrived at the bottom dead center position,the layer developed into a very large cloud of unburnedfuels, which was as long as half an engine stroke length,and as thick as 10 times the piston crevice thickness Notethat the index of the mass fraction contour in Figure 15ranges from 0 to 0.01, while it ranges from 0.0 to 0.06 inFigure 9 Hence, the thickness of the diffusion layer inFigure 15 may appear to be larger than that in Figure 9.From the second plot of Figure 15, which corresponds to

120o CA ATCD, the diffusion layer of unburned HC wasdistributed over a portion of cylinder wall from the middle

of the cylinder wall to the piston position In the next plot,which corresponds to 180o CA ATCD, the bottom end ofthe diffusion layer moved to BDC in accordance with thepiston motion, while the top end of the layer remained at itsprevious position These observations suggests that HCs,which flowed out of the crevice when the piston passed themiddle point of the engine stroke, were not fully oxidized,and an amount of HCs escaped the oxidation process andremained attached at that location In this engine, the pistonpassed the middle of the piston stroke at around 85o CAATCD Therefore, the transition point of unburned HCoxidation in this engine would appear around this crankangle

During the exhaust blow down, the unburned HCs,which were laid down along the cylinder wall, expandedinto the bulk gases in the cylinder In this study, the exhaustvalves were opened at 34o CA BBDC, so some of the HCs

in the diffusion layer at BDC appeared to be entrained bythe rapid motion of the bulk gases, which was generated bythe exhaust blow down This entrainment continued duringthe whole exhaust stroke In the fourth plot of Figure 15,which represents the distribution of the fuel mass fraction

at 120o CA BTDC, unburned HCs on the left side of thecylinder wall were distinctly dragged upward pointing tothe exhaust valves, while those on the right side of the

Figure 14 In-cylinder pressure curves at 1,500 rpm, 0.4 bar

IMAP

Figure 15 Mass fraction of unburned HC during

expan-sion and exhaust stroke

648 H KWON and K MIN

cylinder wall were expanded into the center of the

com-bustion chamber These entrainments of unburned HCs

were more significant at the next plot, which corresponds

to 120o CA BTDC, and some of the HCs on the left side

began to escape from the combustion chamber without

being oxidized In addition, a vortex began to form in the

neighborhood of the piston crevice and cylinder wall

during the exhaust stroke, as shown in the plot of 120o CA

BTDC and 60o CA BTDC in Figure 15 This behavior of

unburned HCs was also observed experimentally by Green

and Cloutman (1997) using planar laser induced fluorescence

(PLIF) In the sequence of pictures by Green and

Clout-man, the wall layer was scraped into a “roll-up” vortex by

the ascending piston, which can also be found in Figure 15

As the piston approached the top dead center position, the

recirculation flow, which was formed in the upper corner of

the cylinder opposite to the exhaust valves, forced the

vortex to be detached from the cylinder wall, and swept by

the bulk gases exiting the combustion chamber In the last

plot of Figure 15, this detachment can be found clearly

However, most of the unburned HCs detached from this

region appeared to be trapped in the combustion chamber

without being swept by exiting bulk gases because the

engine was in part load and low speed condition in the

simulation

Figure 16 shows the mass of unburned HCs in the

cylinder and exhaust port during the expansion and exhaust

strokes As discussed in the results of the 2-D simulations,

the mass of unburned HCs in the cylinder decreased

rapid-ly at the earrapid-ly stage of the expansion stroke due to fast

oxidation, and then the mass of unburned HCs remained

almost constant as the piston passed the transition point In

the figure, a significant change can be seen in the slope of

the curve around 100o CA ATDC After the exhaust valves

were opened, the mass of unburned HCs in the cylinder

decreased continuously during the whole exhaust process

At the early stage of the exhaust process, the mass of

un-burned HCs in the exhaust ports increased rapidly because

of the blow down After that, the flow rate of unburned

HCs from the cylinder to exhaust port was stabilized untilthe late stage of the exhaust stroke, although there weresome fluctuations in the curve As the piston approachedthe exhaust TDC position, the mass of unburned HCs in theexhaust port increased drastically because the unburnedHCs were detached from the cylinder wall and piston topsurface by the recirculation flow, and were forced to flowout of the cylinder by exiting bulk gases From this, it can

be seen that the unburned HCs from the piston creviceswere mainly emitted in the latter parts of the exhaust pro-cess

The resultant oxidation rate of unburned HCs wasevaluated to be 69.8% at EVO, and 79.2% at IVO, whichimplies that 31.1% of additional unburned HCs, escapedfrom the cylinder, was oxidized in the exhaust port Thesevalues are in qualitative agreement with the results ofHamrin and Heywood (1995), who computed that 68% ofthe HCs originated in the combustion chamber crevicesundergo in-cylinder oxidation, and 35% of all HCs exitingthe cylinder undergo exhaust-port oxidation

5 CONCLUSION

In this study, a modified one-step reaction equation wasdeveloped for the modeling of the oxidation of unburnedHCs in SI engines This new equation took the form of anArrhenius’ equation, with modified reaction rate coefficients

to match the 90% oxidation time with the oxidation timeobtained with a detailed chemical reaction mechanism atpractical engine conditions Additionally, the effects of thepressure and intermediate species on the oxidation ratewere investigated over a wide range of variations, and werefound to be of importance, especially at the temperaturerange where the oxidation rate transitioned between effec-tive and ineffective rates As pressure was increased, theoxidation rate increased by up to 40%, and at burnt gasfractions higher than 95%, the oxidation rate increased bymore than 25% These effects were considered in the newone-step reaction equation by scaling the pre-exponentialcoefficient at each pressure and burnt gas fraction condi-tions

Using the new one-step reaction model, the oxidation ofpiston crevice HCs in an SI engine was simulated in a two-dimensional sector mesh, accompanied with a turbulentpremixed combustion model and modified wall function

In the simulations, HCs, which flowed out of the pistoncrevice during the early stage of the expansion stroke, wereoxidized rapidly However, as the in-cylinder temperaturefell below the critical cut-off temperature, which rangedfrom 1,300 to 1,500 K, the oxidation rate became ineffective,and the chemical reaction was almost frozen At the1,500 rpm and 0.4 bar IMAP condition of cold-starting anengine, 44.7% of the total unburned HCs was oxidizedduring the expansion stroke, while 61.5% was oxidized atwarmed-up conditions Engine operational conditions alsoaffected the oxidation rate of the unburned HCs by chang-Figure 16 Mass of unburned HCs in cylinder and in ex-

haust port during expansion and exhaust stroke

MODIFIED ONE-STEP REACTION EQUATION FOR MODELING 649

ing the temperature and pressure of the in-cylinder bulk

gases As the engine load increased from 0.4 to 0.8 bar

IMAP, the oxidation rate was enhanced from 61.5% to

71.3% As the engine speed increased from 1,500 to

3,500 rpm, the oxidation rate was not changed, largely due

to the compromising correlation between the in-cylinder

residence time of the crevice gas and the flow rate out from

the piston crevice The amount of unburned HCs remaining

at EVO was found to be increased as the volume of the

piston crevices increased However, the oxidation rate was

enhanced by increased heat released at the crevice region,

so that the sensitivity of the amount of HCs remaining at

EVO to the piston crevice volume was evaluated to be

linear with a slope of 0.44

Lastly, the one-step reaction model was applied to a

three-dimensional engine mesh to investigate the behavior

of unburned HCs from the piston crevice during the

ex-pansion and exhaust strokes of a conventional engine

During the expansion stroke, the behavior of unburned

HCs was observed to be very similar to the results of the

two-dimensional simulations As the piston reached BDC,

the exhaust valves were opened, and the unburned HC

remaining in the cylinder was forced to flow into the

ex-haust port by a blow down process However, the majority

of the in-cylinder crevice gas was found to be transported

into the exhaust port at the late stage of the exhaust stroke,

when the layers of crevice gases were detached from the

piston and cylinder wall by the recirculation flows

gene-rated in the upper corner of the cylinder The oxidation rate

of the unburned HCs was evaluated to be 69.8% at EVO,

and 79.2% at IVO This result means that 31.1% of the

unburned HCs escaped from the cylinder was oxidized

additionally in the exhaust port

REFERENCES

Alkidas, A C., Drews, R J and Miller, W F (1995) Effects

of piston crevice geometry of the steady-state engine-out

hydrocarbons emissions of a S I engine SAE Paper No

952537

Ando, H., Sakai, Y and Kuwahara, K (2009) Universal

rule of hydrocarbon oxidation SAE Paper No

2009-01-0948

Cheng, W K., Hamrin, D., Heywood, J B., Hochgreb, S.,

Min, K and Norris, M (1993) An overview of

hydro-carbon emissions mechanisms in spark-ignition engines

SAE Paper No 932708

Choi, H., Kim, S and Min, K (2001) Oxidation of

un-burned hydrocarbons from crevices in spark ignition

engines The 4th Int Symp Diagnostics and Modeling of

Combustion in Internal Combustion Engines (COMODIA

2001), Nagoya, Japan

Curran, H J., Gaffuri, P., Pitz, W J and Westbrook, C K

(2002) A comprehensive modeling study of iso-octane

oxidation Combustin and Flame, 129, 253−280

Eng, J A., Leppard, W R., Najt, P M and Dryer, F L

(1997) Experimental hydroarbon consumption rate lations from a spark-ignition engine SAE Paper No.972888

corre-Flowers, D., Aceves, S and Martinez-Frias, J (2003) Effect

of mixing on hydrocarbon and carbon monoxide ssions predictions for isooctane HCCI engine combustionusing a multi-zone detailed kinetics solver SAE Paper

emi-No 2003-01-1821

Green, R M and Cloutman, L D (1997) Planar LIFobservations of unburned fuel escaping the upper ring-land crevice in an SI engine SAE Paper No 970823.Hamrin, D A and Heywood, J B (1995) Modeling ofengine-out hydercaron emissions for prototype productionengines SAE Paper No 950984

Hautman, D J., Dryer, F L., Schug, K P and Glassman, I.(1981) A multiplt-step overall kinetic mechanism forthe oxidation of hydrocarbons Combustion Science and Technology, 25, 219−235

Haworth, D C., Blint, R J., Cuenot, B and Poinsot, T J.(2000) Numerical simulation of turbulent propane-aircombustion with non-homogeneous reactant Combustoin and Flame, 121, 395−417

Heel, B., Maly, R., Weller, H G and Gosman, A D (1998).Validation of SI combustion model over range of speed,load, equivalence ratio and spark timing The 4th Int Symp Diagnostics and Modeling of Combustion in Internal Combustion Engines (COMODIA 98), Nagoya, Japan.Hellström, T and Chomik, J (1995) Oxidation of hydro-carbons released from piston crevices of S.I engine.SAE Paper No 952539

Heywood, J B (1998) Internal Combustion Engine amentals McGraw-Hill Publication New York.Kee, R J., Rupley, F M and Miller, J A (1989) Chemkin II: A fortran Chemical Kinetics Package for The Analysis

Fund-of Gas-Phase Chemical Kinetics Sandia National ratory SAND89-8009.

Labo-Kleemann, A P., Gosman, A D and Binder, K B (2001).Heat transfer in diesel engines: A CFD evaluation study.The 5th Int Symp Diagnostics and Modeling of Com- bustion in Internal Combustion Engines (COMODIA 2001),Nagoya, Japan

Kwon, H., Min, K., Choi, H and Lee, H (2005) Premixedcombustion modeling in an SI engine condering theburned gas composition SAE Paper No. 2005-01-2108.Lee, H B., Kwon, H and Min, K (2007) Effect of variousVVA systems on the engine fuel economy and optimi-zation of a CVVT-VVL SI engine using 1D simulation.Int J Automotive Technology 8, 6, 675−685

Media, S C., Green, R M and Smith, J R (1984) Opticalmeasurements of hydrocarbons emitted from a simulatedcrevice volume in an engine SAE Paper No 840378.Min, K (1994) The Effects of Crevices on the Engine-Out Hydrocarbon Emissions in Spark Ignition Engines Ph

D Dissertation Department of Mechanical Engineering.MIT

Min, K and Cheng, W K (1995) Oxidation of the piston

650 H KWON and K MIN

crevice hydrocarbon during the expansion process in a

spark ignition engine Combustoin Science and Technology,

106, 307−326

Min, K., Cheng, W K and Heywood, J B (1994) The effects

of crevice on the engine-out hydrocarbon emission in SI

engines SAE Paper No 940306

Park, J and Min, K (2000) 3-dimensional modeling of

hydrocarbon oxidation in the exhaust port of a spark

ignition engine Combustion Science and Techonology,

156, 159−171

Sung, C J and Huang, Y (2001) Effect of reformer gas

addition on the laminar flame speeds and flammability

limits of n-butane and iso-butane flames Combustion

and Flame, 126, 1699−1713

Tonse, S R (1996) Numerical simulations of emerging

piston crevices gases SAE Paper No 961968

Weiss, P and Keck, J C (1981) Fast sampling valve

mea-surements of hydrocarbons in the cylinder of CFR aengine SAE Paper No 810149

Weller, H G (1993) The Development of a New Flame Area Combustion Model Using Conditional Averaging.Department of Mechanical Engineering Imperial College

of Science and Technology and Medicine Thermo-FluidsSection ReporTF/9307

Wentworth, J T (1971) The piston crevice volume effect

on exhaust hydrocarbon emission Combustion Science and Technology, 4, 97−100

Westbrook, C K and Dryer, F L (1981) Simplified tion mechanisms for the oxidation of hydrocarbon fules

reac-in flames Combustion Science and Technolohy, 27, 31−

43

Wu, K C and Hochgreb, S (1997) Numerical simulation

of post-flame oxidation of hydrocarbons in spark ignitionengines SAE Paper No 970886