DISI Wall Guided Compression Ignition

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An Investigation of Variable Valve

16V4000G23 ICFN &ICXN

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http://www.paper.edu.cn

Stratified Mixture Formation and Combustion Process for Wall-guided Stratified-charge DISI Engines with Different Piston Bowls by Simulation

2010-01-0595 Published 04/12/2010

Qianwang Fan, Zongjie Hu, Jun Jun Deng and Liguang Li Tongji Univ.

Yi You and Jingyan Hu Geely Automobile Research Institute Copyright © 2010 SAE International

ABSTRACT This paper presents the simulation of in-cylinder stratified mixture formation, spray motion, combustion and emissions in a four-stroke and four valves direct injection spark ignition (DISI) engine with a pent-roof combustion chamber by the computational fluid dynamics (CFD) code. The Extended Coherent Flame Combustion Model (ECFM), implemented in the AVL-Fire codes, was employed. The key parameters of spray characteristics related to computing settings, such as skew angle, cone angle and flow per pulse width with experimental measurements were compared. The numerical analysis is mainly focused on how the tumble flow ratio and geometry of piston bowls affect the motion of charge/spray in-cylinder, the formation of stratified mixture and the combustion and emissions (NO and CO2) for the wall-guided stratified-charge spark-ignition spark-ignition DISI engine. But due to the fuel injected during compression stroke, the effect of intake ports and exhaust ports were not taken into consideration in this study. It is found that the geometry of piston bowls has a major effect on the mixture stratification in-cylinder, the combustion process and others. In addition,

interaction of charge flow, fuel spray, piston bowl as well as combustion.

INTRODUCTION With the increasing attention on achieving substantial improvements of fuel economy and reductions of exhaust emissions, automotive engineers are striving to develop engines with lower Brake Specific Fuel Consumption (BSFC), and which can also comply with future stringent emission requirements. Over the past two decades, many attempts had been made to develop an internal combustion engine for automotive applications that combines the best features of the spark ignition (SI) and the compression ignition (CI) engines. The objective is to combine the specific power of the gasoline engine with the efficiency of the diesel engine at part load. Such an engine would exhibit a BSFC approaching that of the diesel engine, while maintaining the operating characteristics and specific power output of the SI engine [1 [1].The direct injection spark ignition (DISI) engines, in theory, have these two merits. On the one hand, the fuel is injected directly into the combustion chamber in order to have the mixture clouds with an ignitable composition near the spark plug.

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cloud is formed around the spark gap with an overall lean mixture. In reality, the full potential of the DISI combustion system is achieved by utilizing both strategies. At high load the engine should utilize homogeneous charge to maximize air utilization and to avoid soot formation. A slight gain in fuel efficiency is still achieved by charge cooling. Stratified charge is desirable at part load to attain diesel-engine-like fuel economy. At this mode, pumping losses are minimized since engines can operate virtually unthrottled. During the last several years, the DISI engine has been greatly evolved, and various types of DISI productions have been put into the market. Moreover, a lot of investigations have been achieved by experiments and CFD simulations. Especially, with the development of computers and numerical computation methods, in the automotive industry, threedimensional, CFD combustion simulation is increasingly becoming an effective tool for engine design and development. In the past, the application of transient CFD incylinder analysis has mainly been focused on the intake and compression strokes, calculating in-cylinder transient tumble curves for an assessment of probable combustion stability, and on air-fuel mixture formation for combustibility [3, 4]. Luca Olmo [5] presented that the in-cylinder flow and combustion analysis of a high performance four valves DISI engine has been successfully carried out for three different design configurations, adopting the ECFM combustion model implemented in the STAR-CD code. Jianwen Yi et al [ 6] studied the interaction between in-cylinder flow and injected fuel spray in DISI engines, the results showed that the major effect of intake flow on the fuel spray is that the induced flow tends to make the spray collapse. Sungjun Kim et al [ 7] investigated the air-fuel mixture formation and combustion characteristics in spray-guided DISI engine employing a StarCD code. The results showed the enhanced tumble flow can deteriorate the mixture distribution and decrease the burning rate. Moreover, G. Fontana et al [ 8] studied the effect of the different combustion chambers on engine performances and emissions of a small gasoline engine, employing numerical and experimental methods. Although previously a lot of investigations had been performed, including stratified-charge formation [9,10] and the effect of various parameters on engine performances and emissions [11,12] etc. The objectives of this paper are to

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ENGINE SPECIFICATIONS

Engine Specifications Each cylinder of the four-cylinder engine which is employed has a pentroof head, four valves (two intake valves and two exhaust valves), a centrally mounted spark plug, a piston with alternative bowls (shown in Fig.1), and an six-hole injector mounted under intake port at the inclination angle of 28 degrees with respect to the cylinder head, whose holes are not located evenly (shown in Fig.2). The Fig.1 shows two pistons, piston A and piston B. In addition, in order to adapt each piston bowl shape, the cylinder head was also modified slightly, and the injector installation exists in difference, namely, for piston A, the maximal included angle between two nozzle holes is located near the side of the cylinder head, but for piston B, one located near the side of piston top. For the injector, the spray characteristics related to computation later, were given based on the experimental data, which are shown in Fig 3 and Table 1. In addition, the engine specification is summarized in Table 1 as well.

(a). Scheme of piston A

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Table 1. Engine specification and injector parameters

Fig. 2. The scheme of nozzle holes' distribution

Note: 0 Degree Crank Angle is assigned in TDC intake stroke

BASIC THEORY AND MODEL The CFD calculations were performed with AVL Fire codes. The codes solve ensemble-averaged equations for momentum, energy, species concentration and mass. The calculations start from the time of intake valve close (IVC 220°CA ATDC) to the time of exhaust valve open (EVO 510°CA ATDC), calculating step lengths are 0.2°CA during fuel injection and combustion, others are 0.5°CA. During the whole process, the flow motion event was computed using the turbulence model, the spray and combustion events were performed using Discrete Droplet Model (DDM) and Extended Coherent Flame Model (ECFM). In addition, equations of energy, motion and mass were solved using Upwind Scheme, Central Differencing Scheme and MINMOD Relaxed Scheme [13,14,16], respectively. The theories associated with the models were to be simply introduced below, for details, referring to the literatures [13,14,15,16].

Fig.3. the scheme of spray event in ambient conditions

Mesh In the study, the unstructured grids were adopted, while the piston was modeled by the dynamic mesh models. The entire process simulating piston movement was achieved by the dynamic grids, which were completed using FAME Meshing (FM) and FAME Engine Plus (FEP) tools. Both of models, at

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Turbulent flow model The k-ε model is used as turbulence model. Here, the k equation [13, 15] and ε equation [13, 15] are expressed respectively, as follows:

(1) Fig.4. The scheme of piston A

Where,

is the turbulent kinetic energy,

is the tensor of

strain rate associated to fluctuating flow; is the tensor of strain rate associated to mean flow; is kinematic viscosity coefficient.

(2) Where, Fig.5. The scheme of piston B

Initial Conditions and Boundary Conditions Setting The initial conditions and boundary conditions are provided resulting from computational values of GT-Power code, and some referred experimental values. Table.2 shows essential values of initial and boundary conditions. Table.2. Initial Conditions and Boundary Conditions Setting

is the density,

Prandtl number of

,

is the dissipation rate,

is the

is the effective viscosity

coefficient, , are the constant, kinetic energy production term.

is the turbulent

Combustion model The Extended Coherent Flame Model (ECFM) [ 13, 15] has been mainly developed in order to describe combustion in DISI engines. This model is fully coupled to the spray model and enables stratified combustion modeling including EGR effects and NO formation. The model relies on a conditional unburnt/burnt description of the thermochemical properties of the gas. The ECFM contains all the features of the standard CFM and the improvements of the MCFM. Differences to the other coherent flame models are described in the references [13, 15]. For turbulent combustion phenomena, the ECFM model leads to the calculation of the mean fuel reaction rate. Hence, this model uses a 2-step chemistry mechanism [13] for the fuel conversion like:

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The mean turbulent fuel reaction rate [13] is computed as the product of the flame surface density Σ and the laminar burning velocity SL via:

(5)

Where, is the stoichiometric coefficients of species in the reaction r, while for the reactants these coefficients are negative and for the products positive, respectively.

of the liquid fuel, gas enthalpy.

is the mass fraction,

is the fresh

Spray model Spray simulations involve multi-phase flow phenomena and as such require the numerical solution of conservation equations for the gas and the liquid phase simultaneously. With respect to the liquid phase, practically all spray calculations in the engineering environment today are based on a statistical method referred to as the Discrete Droplet Method (DDM) [13, 15]. Movement equation

is

is the mean

laminar fuel consumption rate described earlier. function depending on the equivalence ratio carbon and hydrogen atoms, respectively.

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whereafter, some basic equations are introduced below.

the reaction rate for reaction (3) and (4).


is a (9)

, number of Energy equation (heat transfer)

From the previous sections it is obvious that the extended CFM can be closed if the local properties of the burnt and unburnt gases are known. Hence, in each computational cell two concentrations are calculated: a concentration in the unburnt gases and a concentration in the burnt gases, respectively. Hence two additional transport equations have to be introduced, one for the unburnt fuel mass fraction and one for the unburnt oxygen mass fraction. Below the two transport equations [13] are given.

(10) Mass equation (evaporation)

(11) (6)

Where,

is the Particle velocity vector [m/s],

Particle diameter [m], is the gas density [kg/m3], the droplet velocity [m/s]; (7) Additionally, a transport equation for the unburnt gas enthalpy is also introduced as shown below.

is the droplet density [kg/m3], is the gas velocity [m/s],

is

is the Particle mass [kg],

is the Specific heat of liquid [J/(kg.K)], temperature [K],

is the

is the droplet

is the particle surface temperature [K],

is the Nusselt number, is the Particle far-field temperature [K], λ is the Thermal conductivity [W/(m.K)].

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timing of 345 °CA ATDC, by varying tumble flow ratio for two different piston bowl shapes.

Interaction between flow and spray Fig.6 shows the interaction between charge flow and spray inside combustion chamber for the stratified-charge directinjection gasoline engine. When the spray is induced into cylinder, in-cylinder single vortex is transformed two ones around the spray for each piston. Before the spray impinges the piston bowl surface, many droplets change their directions and move perpendicularly to the spray axis toward the top of the combustion chamber. The reason for this drastic change in the flow direction is the strong influence of the in-cylinder airflow on the smaller droplets [17]. Then, at the end of injection, the spray head direction is changed along piston bowl surface for each piston, due to the effect of piston bowl profile and tumble flow. As far as two pistons are concerned, the piston B seems to be more suitable for mixture stratification at the injection timing of 310 °CA ATDC.
In addition, different tumble intensity causes the spray velocity difference along the piston bowl contour for two piston bowls. Moreover, there exists a distinct spray plume structure inside combustion chamber with two different pistons, mainly resulting from asymmetric distribution of holes, and as mentioned before, the nozzle holes are located differently around the axis of the hole where the injector is installed.

In-cylinder mixture strength The mixture formation has a significant impact on combustion and emissions. Therefore, it is quite important to study in-cylinder fuel-air distribution. The references [3, 4, 5, 10] presented the results of mixture formation. In those investigations, the spatial distribution and temporal history of mixture formation were studied. Besides, mixture strength is designated as equivalence ratio. For an engine operation in the stratified charge mode, the spatial distribution and temporal development of the mixture strength are shown in Fig. 7. At the end of fuel injection (324 °CA ATDC), the thickest mixture exists inside the downright side of combustion chamber, due mainly to the incomplete evaporated spray droplets. As the piston moves



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the time of 310 °CA ATDC. At the same time, as shown in Fig.8(d), it is also acquired, for the piston B under tumble flow ratio (TR) of 1, that the appropriate mixture around the spark gap is formed, resulting mainly from the effect of intensified tumble flow.

Combustion Characteristics Fig.8 shows the spatial distribution and temporal history of in-cylinder temperature throughout the combustion chamber. Firstly, it shows that the spatial distribution of in-cylinder temperature for the piston bowl B is superior to that of the piston bowl A, which is the result of favorable stratifiedmixture formation for the piston bowl B. Furthermore, it can be also found that tumble intensity has a strong influence on the in-cylinder temperature. Lastly, it implies that the stratified mixture leads to in-cylinder temperature stratification throughout the combustion chamber, especially for the piston bowl B, it is more obvious as the result of appropriate stratified mixture and stable combustion.

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affects flame kernel formation and flame propagation, hence, with the tumble flow ratio of 1, in-cylinder combustion process is satisfying, due mainly to appropriate tumble flow intensity improving mixture formation and accelerating robust combustion. For the piston A, when the tumble flow ratio is the value of 0.5, there exist two distinct peak values of mean in-cylinder temperature. The second peak value arises resulting mainly from the strong post-combustion.

Emissions

Fig.10. Effect of piston bowl on mean in-cylinder temperature at various tumble flow ratios

Fig.9 shows the comparisons of mean in-cylinder pressure, rate of heat release and accumulated heat release between the piston A and the piston B under different tumble flow ratio (TR). It can be found that the mean pressure is some slightly different between the piston A and the piston B. However, the strong tumble flow intensity leads to mean pressure increase due to stronger tumble flow improving mixture formation and accelerating combustion process, but tumble flow is too strong to form proper air-fuel mixture in the position of spark plug, so tumble intensity is enhanced properly. Furthermore, as far as the accumulated heat release in Fig.9 concerned, compared to the piston B, the piston A is sensitive to tumble flow intensity, due to piston A shape resulting in the worse air-fuel mixture with the TR of 0.5. Hence it is clearly effective for the piston A with the TR of 1. For the piston B, the rate of heat release and the accumulated heat release are higher than that of the piston A, regardless of tumble intensity. Moreover, the strong tumble flow causes them to rise greatly. The main reasons are that the fast burning period becomes shorter, and that the location of area center in the curve of heat release rate is closer to the top dead center (TDC) and the combustion process is remarkably improved [18].

Fig.11 shows the effect of piston bowl types on CO 2 emissions at various tumble flow ratios. Regardless of tumble flow ratio, for the piston B, CO 2 emissions, are generated more than that of the piston A, due to mainly appropriate stratified-charge mixture formation and relatively complete combustion. On the other hand, regardless of piston bowl shape, the stronger tumble flow promotes more appropriate stratified-charge mixture formation and more adequate combustion, leading to more CO2 emissions generation. Combined with the accumulated heat release in Fig.9, it is higher for the piston A than that of the piston bowl B with the TR of 1. However, it shows that carbon dioxide is less for the piston A, as its piston bowl shape resulting in different airfuel mixture distribution. Although it is clearly effective for air-fuel mixture formation of the piston bowl A to enhance tumble flow, but compared to the piston B, air-fuel mixture stratification is worse for the piston A, so a lot of carbon monoxide emission is generated. Besides, the unburnt hydrocarbon is more for the piston B due to better air-fuel mixture stratification resulting in quenching in the position far from the spark plug. Fig.12 shows the effect of piston bowl types on NO emissions formation at various tumble flow ratios. Generally speaking, the NOx formation strongly depends on the combustion temperature and combustion duration at high temperature. Likewise, for piston B, NO emissions are generated more than that of the piston A, regardless of tumble flow ratio, due to mainly the piston B fitting more to stratified-charge mixture formation and robust combustion. On the other hand, for each piston bowl, the stronger tumble flow is formed, the more amount of NO emissions is generated. Moreover, for piston A at tumble flow ratio of 0.5,

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However, as tumble flow ratio increases, heat release rate is enhanced, and the phase is advanced, as the result that the stronger tumble flow causes beneficial air-fuel mixture around the spark gap, and better stratified-charge mixture throughout combustion chamber. In addition, accumulated heat release also increases, but there are slightly distinct variations for mean in-cylinder pressure. Fig.14 shows the effect of different tumble flow ratios on mean in-cylinder temperature. It shows that mean in-cylinder temperature dramatically increases as tumble flow is enhanced. In addition, the inflexion of mean temperature is advanced as well as it is possibly attributed to tumble flow improving in-cylinder mixture distribution and enhancing turbulence flow near top dead center (TDC). Finally, all accelerate the rate of heat release rate. Fig.11. Effect of piston bowl on CO 2 emissions

Fig.12. Effect of piston bowl on NO emissions

The effect of tumble flow on combustion and emissions for piston B It is known that in-cylinder airflow fields have the prominent influence on combustion and emission characteristics. Therefore, the section presents the effect of the tumble intensity on mean in-cylinder pressure, the rate of heat release

Fig. 13. Effect of tumble flow on combustion

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Fig.14. Effect of tumble flow on mean in-cylinder temperature


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Fig. 16. Effect of tumble flow ratio on nitrous oxide

The CO2 and NO emissions are given in Fig.15 and Fig.16, respectively. In Fig.15, it can be found that, increasing tumble flow ratio, the starting of combustion is slightly advanced, and the amount of CO2 rises dramatically as well, due mainly to appropriate stratified-charge mixture caused by tumble flow. The NO formation depends strongly on in-cylinder temperature and combustion duration at high temperature. The results shown in Fig.16 keep in accordance with the principle. The stronger tumble flow promotes the better stratified-charge formation, leading to a stable combustion generation.

Fig. 15. Effect of tumble flow ratio on carbon dioxide

The effect of injection timing on combustion and in-cylinder emissions for piston bowl B at TR=0.5 Combustion The start of injection (SOI) is an important parameter because it affects the combustion characteristics and exhaust emissions of the engine. To achieve the higher degree of stratification of the air-fuel mixture, the accurate control of the quantity and timing of fuel injection is necessary. Therefore, it is important to investigate the effect of injection timing on combustion and emission characteristics later. In

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combustion dramatically improves combustion process at part load, it can be concluded that the delay of injection timing can lead to the desired stratification of the mixture. Hence, it can be found that, the appropriate delay of injection timing causes mean in-cylinder temperature to rise.

Emissions In this part of the study, the exhaust emissions such as CO 2, NO, were compared for the three injection timings at the different crank angles. The relationships between each emission and injection timing were separately discussed. Some observations between the emissions were also given in this paragraph.

Fig.17. Effect of injection timing on combustion

Fig.19. Effect of injection timing on carbon dioxide

Fig.18. Effect of injection timing on mean temperature

Fig.17 shows the mean in-cylinder pressure, heat release rate and accumulated heat release characteristics. As injection timing is retarded, in-cylinder combustion characteristics don't vary monotonously, in other words, there exists an optimization value of injection timing. In the calculating

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CONTACT

In addition, other literatures [19] present some information, which fuel stratification within the cylinder produces local equivalence ratio different from the global equivalence ratio. Therefore, the local fuel enrichment will result in relatively high local temperatures. Thus, the high local temperatures provide the energy required to form NO at a rapid rate. Fuel stratification can be achieved by retarded SOI timing in a direct injection system.

REFERENCES

The simulating results of this study were summarized as follows: The piston bowl shape is a key parameter, which can promote the stratified-charge formation and combustion process in the SCDI gasoline engine. In this case, compared to the piston A, the piston B is more suitable for the stratified combustion at the injection timing of 310°CA ATDC, subsequently, the ignition timing of 345°CA ATDC. In general, tumble flow affects the degree of mixture stratification, flame kernel formation and flame diffusion. The stronger tumble flow exists in the combustion chamber, the better stratified-charge mixture is formed, and however, tumble flow is excessively strong resulting in blowing out the flame kernel, misfiring and quenching and so on. To achieve high degree of stratification of the charge, the accurate control of injection timing is necessary. Advancing or retarding injection timing excessively, deteriorate mixture formation, therefore, there is an optimization value. For the investigated piston B, the injection timing of 298°CA ATDC is more appropriate at ignition timing of 345°CA ATDC with

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In general, at the temperatures over 1800 K, NO generation rate increases rapidly with increasing combustion temperature. Therefore, the peak combustion temperature must be kept below 1800 K to prevent the rapid NO formation. At the injection timing of 298 °CA ATDC, NO emissions generate relatively high, as the result of adequate combustion leading to a higher temperature and combustion period at high temperature. At the injection timing of 278°CA ATDC, because the stratified-charge mixture is formed unsuitably, especially at the time of ignition, combustion temperature can't reach the level of lots of NO generated, resulted the NO emissions at the level of nearly to zero.

CONCLUSIONS


Dr Liguang Li, Prof. School of Automotive Studies Tongji University, Shanghai, 201804, P.R. China. [email protected] Tel: +86 21 69583817 Fax: +86 21 69589978 Qianwang Fan PhD Candidate [email protected]

1. Zhao, F., “Automotive Gasoline Direct Injection Engines,” SAE International, Warrendale, PA, ISBN 978-0-7680-0882-1, 2002. 2. Rotondi Rossella, Bella Gino. Gasoline direct injection spray simulation. International Journal of Thermal Sciences 45 (2006) 168-179. 3. Schänzlin, K., Koch, T., Tzannis, A.P., and Boulouchos, K., “Characterization of Mixture Formation in a Direct Injected Spark Ignition Engine”, SAE Technical Paper 2001-01-1909, 2001. 4. Hélie, J., Duclos, J.-M., Baritaud, T., Poinsot, T. et al., “Influence of Mixture Fluctuations on Combustion in Direct Injection Spark Ignition Engines Simulations,” SAE Technical Paper 2001-01-1226, 2001. 5. Olmo, L. and Thornton, J., “CFD Analysis of Mixture Formation and Combustion Process for High Performance DI Gasoline Engine,” SAE Technical Paper 2005-01-0214, 2005. 6. Yi, J., Han, Z., Yang, J., Anderson, R. et al., “Modeling of the Interaction of Intake Flow and Fuel Spray in DISI Engines,” SAE Technical Paper 2000-01-0656, 2000. 7. Kim, S.-J., Kim, Y.-N., and Lee, J.-H., “Analysis of the In-Cylinder Flow, Mixture Formation and Combustion Processes in a Spray-Guided DISI Engines,” SAE Technical Paper 2008-01-0142, 2008. 8. Fontana, G., Galloni, E., Palmaccio, R., and Torella, E., “Numerical and Experimental Analysis of Different Combustion Chambers for a Small Spark-Ignition Engine,”

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11. Kaiser, E. W., Siegl, W.O., Brehob, D.D., and Haghgooie, M.,“Engine-Out Emissions from a DirectInjection Spark-Ignition (DISI) Engine,” SAE Technical Paper 1999-01-1529, 1999. 12. Muñoz, R.H., Han, Z., VanDerWege, B., and Yi, J., “Effect of Compression Ratio on Stratified-Charge DirectInjection Gasoline Combustion,” SAE Technical Paper 2005-01-0100, 2005.



DI

Direct Injection ECFM Extended Coherent Flame Model FM

FAME Meshing

13. AVL Fire_2008_CFD User Manual. 14. Anderson John D.. Computational Fluid Dynamics. 15. Xie Maozhao. Computational Combustion of Internal Combustion Engine. Dalian University of Science and Technology press, China, 2005.

FEP

16. Wu Deming and Hao Ye. Practical Computational Fluid Dynamics Fundamentals. Haerbin Engineering Universuty press, China, 2006.

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17. Hentschel Werner. Optical diagnostics for combustion process development direct injection gasoline engines. Proceedings of the Combustion Institute 28:1119-1135, 2005.

NO

18. Wei Li, Ying Wang et al. Study on improvement of fuel economy and reduction in emissions for stoichiometric gasoline engines. Applied Thermal Engineering 27 2919-2923, 2007.

PFI

19. Canakci Mustafa. An experimental study for the effects of boost pressure on the performance and exhaust emissions of a DI-HCCI gasoline engine. Fuel 87:1503-1514,2008.

FAME Meshing Plus

Direct Injection Spark Ignition

Nitrogen oxide

Port Fuel Injection SCDI

Stratified Charge Direct Injection SOI

Start of Injection

ABBREVIATION AND DEFINITION ATDC After Top Dead Center BSFC

Brake Specific Fuel Consumption BC

Boundary Condition CA

Crank Angle CFD

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TR

Tumble flow ratio

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The Engineering Meetings Board has approved this paper for publication. It has

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Positions and opinions advanced in this paper are those of the author(s) and not

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