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Effect of piston bowl geometry on flow, combustion and emission in DI diesel engine--a CFD approach.

Introduction

Internal combustion engines have been a relatively inexpensive and reliable source of power for applications ranging from domestic use to large scale industrial and transportation applications for most of the twentieth century. Their dependability, along with a seemingly inexhaustile fuel source, led to their widespread acceptance within a few decades of their introduction. Recently, however, heightened concern over the environmental impact of internal combustion engines has led to increasing governmental regulations regarding the emissions and fuel economy performance of both compression ignition (diesel) and spark ignition internal combustion engines.

DI Diesel engines, having the evident benefit of a higher thermal efficiency than all other engines, have served for both light- duty and heavy-duty vehicles. However direct injection diesel engines emit more particulates and oxides of nitrogen than their counterpart and hence a reduction of such emissions is most urgent. Unfortunately the tendencies of the particulates and oxides of nitrogen emissions are contradictory, and it is not easy to keep them within the level of the coming stringent regulations.

To obtain a better combustion with lesser emissions in direct--injection diesel engines, it is necessary to achieve a good spatial distribution of the injected fuel throughout the entire space [1]. This requires matching of the fuel sprays with combustion chamber geometry to effectively make use of the gas flows. In other words, matching the combustion chamber geometry, fuel injection and gas flows is the most crucial factor for attaining a better combustion [4].

In DI diesel engines, swirl can increase the rate of fuel-air mixing [3], reducing the combustion duration for re-entrant chambers at retarded injection timings. Swirl interaction[8] with compression induced squish flow increases turbulence levels in the combustion bowl, promoting mixing. Since the flow in the combustion chamber develops from interaction of the intake flow with the in-cylinder geometry, the goal of this work is to characterize the role of combustion chamber geometry on in-cylinder flow, thus the fuel-air mixing, combustion and pollutant formation processes.

From the literatures, authors found that the effect of geometry has a negligible effect on the airflow during the intake stroke and early part of the compression stroke. But when the piston moves towards Top Dead Centre (TDC), the bowl geometry has a significant effect on air flow thereby resulting in better atomization, better mixing and better combustion. The re-entrant chamber without central projection and with sharp edges provides higher swirl number than all other chambers. This higher swirl number reduces the soot emission at the cost of higher [N.sub.OX] level. The increase in [NO.sub.X] can be reduced by retarded injection timing with high-pressure injection. Higher power can be achieved by using boost pressure system (turbo charging) with high pressure (CRDI) split- injection.

Numerical Methodology

In order to study the flow characteristics, CFDcalculations of the suction and compression strokes have been performed. The pre-processor GAMBIT is used to create the computational domain of the engine and commercial computational fluid dynamics code STAR-CD is used for solution of governing equations and post processing the results. A hexahedral block structured mesh was employed for the entire computational domain of the engine with 369785 cells. A finite volume code, STAR-CD, has been used to solve the discretized Navier--Stokes equations. The RNG k--[epsilon] model with standard wall function is used. The program is based on the pressurecorrection method and uses the PISO algorithm. The first order upwind differencing scheme (UD) is used for the momentum, energy and turbulence equations and the temporal discretization is implicit.

The calculations begin at TDC of the intake stroke and finish 30 CAD after TDC of compression. Constant pressure boundary conditions were assigned for both intake and exhaust ports, so the dynamic effects were neglected. The initial values for pressure and temperature were 1.02 bar and 303 K respectively, with both variables considered as homogeneous in the whole domain. As the residual swirl of the flow in the cylinder at the end of the exhaust stroke was not taken into account, the flow was supposed to be quiescent initially. The initial turbulence intensity was set at 5% of the mean flow, and the integral length scale was estimated with the mixing length model of Prandtl. The walls of the intake ports, the lateral walls of the valves and the cylinder head, the cylinder wall and the piston crown that form the walls of the combustion chamber were considered adiabatic.

Case Description

The engine studied in this work is a stationary, single-cylinder direct-injection Diesel engine with three different piston shapes were considered. These shapes are representative of the geometries usually employed for the optimum combustion process in real engines. The piston named Bowl A has a spherical chamber and used as a baseline model . The piston named Bowl B is a Toroidal chamber The piston named Bowl C is a largely reentrant geometry. The engine specifications are given in Table 1.

[FIGURE 1 OMITTED]

[FIGURE 2 OMITTED]

[FIGURE 3 OMITTED]

Results and Discussion

Motoring Case Result

The flow inside the cylinder during the intake and compression stroke is analyzed in this section. In order to study the averaged transverse fluid flow behaviour particularly to be used in the analysis,a quantity known as swirl ratio (SR) is computed for the in-cylinder flow field about the cylinder axis, and is expressed as,

SR = 60 [H.sub.z]/2[pi][M.sub.z][[OMEGA].sub.cs]

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII.]

[MATHEMATICAL EXPRESSION NOT REPRODUCIBLE IN ASCII.]

where Hz represents the total angular momentum of the in-cylinder fluid about the cylinder axis

Mz is the total moment of inertia of the fluid about the cylinder axis,

[OMEGA]cs is the angular speed of the crankshaft in rpm,

[u.sub.i] and [v.sub.i] represent the local components of velocity in the x and y directions respectively and mi, the mass accumulated in each cell.

[FIGURE 4 OMITTED]

Fig4 presents the temporal distribution of mass averaged swirl ratio during suction and compression The swirl is generated early during the intake stroke and the maximum value is reached at 90 CAD aTDC at which point the piston reaches its maximum instantaneous speed. After this, the discharge velocity into the cylinder decreases and swirl drops slowly during the rest of the intake stroke. This reducing trend continues in the first part of the compression stroke due to friction at the wall. Early during the first stage of compression stroke, between BDC and 250 cad aTDC the stratified swirl structure obtained at the end of the intake stroke is maintained. However, as the compression advances,In the second stage between 250 cad aTDC and TDC the axial upwards flow induces a gradual increase of the swirl velocity in the top part of the piston.

However, when piston approaching TDC, swirl is enhanced as the flow accelerates in preserving its angular momentum within the smaller diameter piston bowl. During the expansion phase, reverse squish as the flow exits from the piston bowl and wall friction contribute to the sudden fall of the swirl velocity.

Bowl B has a larger diameter and consequently, swirl is significantly smaller around TDC, as can be seen in Fig.4. Near to TDC, the flow is confined in the piston bowl and as a result, the angular velocity increases. Since the bowl diameter of Bowl B is larger than that of Bowl A, the confinement effect is less important at the beginning of this phase and the angular velocity increase is slower and largely mitigated by the losses induced by wall friction effects. Indeed, for Bowl B there is even a slight reduction of the swirl ratio between 250 cad aTDC and TDC The highest swirl ratio is obtained with Bowl C as it has the smallest entry radius. It is observed that ombustion chamber shape has a negligible effect on in-cylinder flow dynamics during initial part of the compression. But near TDC of compression stroke is an important period with respect to combustion and pollutants formation,

Combustion case Results

The fuel taken for this analysis is n-dodecane (C12H26), which is considered as an equivalent hydrocarbon to diesel fuel for computational research. Properties of ndodecane are incorporated in CFD code. For simulating combustion, first a chemical reaction scheme is defined, which is a single step global reaction. A proper reaction rate mechanism based on eddy break-up model is also specified. Based on the shell auto ignition model, the combustion is initiated and it proceeds according to the eddy brake up model. Various model used for combustion simulation are reported in table.2

The combustion and emission related parameters presented in this section are

* Pressure and temperature variations inside the cylinder

* [NO.sub.X] and soot distribution inside the cylinder

Pressure and Temperature variations inside the cylinder

Figure5 shows that Bowl C has the highest peak pressure and Temperature of all the three bowls and hence, the lowest SFC. The difference between the peak pressure values of Bowl C and Bowl A bowl is 4.5 bar and it is due to strong air motion induced in Bowl C increases the rate of combustion. Better combustion in turn results in higher level of pressure and temperature in Bowl C

[FIGURE 5 OMITTED]

[NO.sub.X] and Soot emission

Bowl C with its high swirl ratio (observed in Figure 4) exhibits superior soot oxidation (Figure 6). In this case, two different mechanisms draw responsible, namely a promoted fuel-air mixing process in the early phase of combustion and hence reduced soot production and a higher turbulent mixing intensity leading to higher local oxidation rates in the late combustion phase. It is observed that there is reduction of soot emission in Bowl C than Bowl A and Bowl B

Thermal [NO.sub.X] formation at the beginning of injection is rapid for the all the three cases. This is due to the high temperature rise during combustion of fuel. The effect of turbulence increases the rate of mixing of fuel and air resulting in better combustion and higher temperature. This higher temperature tends to increase [NO.sub.X] production as shown in Figure 8. As a result, the [NO.sub.X] emission of Bowl C is more than Bowl A and Bowl B

[FIGURE 6 OMITTED]

Conclusion Based on this investigation, the following conclusions are drawn:

* Bowl C creates more swirl than baseline bowl Bowl A and Bowl B at firing TDC

* The peak pressure and temperature are obtained at 4[degrees] CA aTDC for all the three cases. The pressure and temperature obtained are higher for Bowl C

* Reduction in soot is achieved for the Bowl C compared to bowl A and Bowl B

* However, this is at the expense of [NO.sub.X] which increases by about 15 % due to high temperature rise in Bowl C

* Bowl B was found to be worse than the baseline bowl based on the performance and emission,

In summary, Bowl C enhance the turbulence and hence results in better air-fuel mixing process among all three bowls in a DI diesel engine. As a result, the ISFC and soot emission are reduced, although the [NO.sub.X] emission is increased owing to better mixing and a faster combustion process. Globally, since the reduction of soot is larger than the increase of [NO.sub.X] it can be concluded that Bowl C is the best trade-off between performance and emissions. The accurate prediction of pressure traces and heat release data is what is ultimately desired to allow simulation of potential bowl designs without having to proto-type every possible design variation. This capability allows more potential designs to be evaluated to identify the promising few for prototyping and testing. This will lead to substantial savings of time and money in developing more sophisticated designs of diesel engines.

References

[1] Arturo de Risi, Teresa Donateo, Domenico Laforgia,"Optimization of the Combustion Chamber of Direct Injection Diesel Engines" SAE 2003-01-1064.

[2] Bianchi, G.M., Pelloni, P., Corcine, F.E. Mattarelli, E.Luppino, Bertoni. F, " Numerical Study of the Combustion Chamber Shape for Common Rail H.S.D.I Diesel Engines", SAE 2000-01-1179.

[3] Corcione. F. E, Annunziata Fusca, and Gerardo Valentino, "Numerical and Experimental Analysis of Diesel Air Fuel Mixing" SAE 931948.

[4] Herbert Schapertons, Fred Thiele," Three Dimensional Computations for Flow Fields in D I Piston Bowls". SAE 860463.

[5] Ikegami. M, Fukuda. M, Yoshihara. Y and Kaneko. J, "Combustion Chamber Shape and Pressurized Injection in High- Speed Direct--Injection Diesel Engines" SAE 900440.

[6] Lu Lin, Duan Shulin, Xiao Jin, Wu Jinxiang and Xiaohong, "Effect of Combustion Chamber Geometry on In-Cylinder Air Motion and Performance in DI Diesel Engine", SAE2000-01-0510.

[7] Montajir. R, "New Combustion Chamber Concept for Low Emission Small DI Diesel Engine", 2001-01-1763.

[8] Ogawa. H, Matsui. Y, Kimura. S, Kawashima. J, "Three Dimensional Computation of the Effects of the Swirl Ratio in Direct- Injection Diesel Engines on [NO.sub.X] and Soot emissions" SAE 961125.

[9] Payri. F, Benajes. J, Margot. X, Gil. A, "CFD Modeling of the In-Cylinder Flow in Direct --Injection Diesel Engines" Computers and Fluids 33 (2004) 995-1021.

A. Gunabalan and R. Ramaprabhu Department of Mechanical Engineering, Anna University, Chennai
Table 1: Engine specifications.

Number of cylinders     One

Bore                    87.5 mm
Stroke                  110 mm
Connecting rod length   232 mm
Compression ratio       17.5:1
Engine speed            1500 mm
Power                   6 bhp
Injection pressure      230 bar
Start of Injection      23 deg
Injection Duration      30 deg
Number of holes         3

Table 2: Models used for combustion Simulation.

S. No   ParameteModel used

1       Turbulence Model         RNG k - [epsilon] Model
2       Combustion Model         EBU-LATCT Model
3       Ignition Model           Shell Auto-ignition Model
4       Nozzle Flow Model        Effective Nozzle Model
5       Droplet Brake-up Model   Reitz--Diwakar Break-up Model
6       Atomization Model        Huh Atomization Model
7       Collision Model          O'Rourke Collision Model
8       NOx Model                Extended Zeldovich Mechanism
9       Soot Model               Mauss Soot Model
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Author:Gunabalan, A.; Ramaprabhu, R.
Publication:International Journal of Applied Engineering Research
Article Type:Report
Geographic Code:1USA
Date:Nov 1, 2009
Words:2305
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