Control of peak pressures of an HCCI engine under varying swirl and operating parameters

Abstract

The major advantages of homogeneous charge compression ignition (HCCI) are high efficiency in combination with low NO_x-emissions. However, one of the major challenges with HCCI is the control of higher peak pressures which may damage the engine, limiting the HCCI engine life period. In this paper, an attempt is made to analyze computationally the effect of induction swirl in controlling the peak pressures of an HCCI engine under various operating parameters. A single cylinder 1.6 L reentrant piston bowl diesel engine is chosen. For computational analysis, the ECFM-3Z model of STAR –CD is considered because it is suitable for analyzing the combustion processes in SI and CI engines. As an HCCI engine is a hybrid version of SI and CI engines, the ECFM-3Z model with necessary modifications is used to analyze the peak pressures inside the combustion chamber. The ECFM-3Z model for HCCI mode of combustion is validated with the existing literature to make sure that the results obtained are accurate. Numerical experiments are performed to study the effect of varying properties like speed of the engine, piston bowl geometry, exhaust gas recirculation (EGR) and equivalence ratio under different swirl ratios in controlling the peak pressures inside the combustion chamber. The results show that the swirl ratio has a considerable impact on controlling the peak pressures of HCCI engine. A reduction in peak pressures are observed with a swirl ratio of 4 because of reduced in cylinder temperatures. The combined effect of four operating parameters, i.e., the speed of the engine, piston bowl geometry, EGR, and equivalence ratio with swirl ratios suggest that lower intake temperatures, reentrant piston bowl, higher engine speeds and higher swirl ratios are favorable in controlling the peak pressures.

Keywords

HCCI engine / ECFM-3Z / Swirl ratio / peak pressures / engine speed / piston bowl geometry

Introduction

The homogeneous charge compression ignition (HCCI) is a promising alternative for combustion in the internal combustion engine. The HCCI concept is a hybrid of the successful spark ignition (SI) and compression ignition(CI) engine concepts. In a diesel engine, the fuel is exposed to high enough temperature for auto ignition, but in contradiction to the diesel engine type, a homogeneous fuel/air-mixture is used. The homogeneous mixture is created in the intake system, using a low pressure injection system. To limit the rate of combustion, much diluted mixtures must be used. HCCI is a concept of hybrid combustion, between Otto engine and diesel engine. HCCI is, however, not a modern finding. Already in the early twentieth century, hot bulb engines operated with an HCCI-like combustion. They were superior in terms of brake efficiency compared with the contemporary gasoline engines and at the same level as the diesel engines. High engine efficiencies, ultra low NO emissions, and low particulates are the benefits of HCCI engines. Volumetric auto ignited combustion of compressed lean air-fuel mixture is attributed to these benefits. There are also a few drawbacks in HCCI engines, like low specific output, narrow operating range, lack control over the ignition process, long start up time and high emissions like CO and UHC emissions. The CO and UHC emissions can be after treated using catalytic converters [ 1– 3]. Heywood [ 4] explained the combustion characteristics of HCCI combustion along with the many other fundamental concepts about the combustion in an IC engine. Heywood [ 4] reported that the problems like high particulate matter and soot emissions (because of fuel rich and diffusion rich regions) associated with the conventional CI engine could be overcome by HCCI engines.Volumetric combustion and low temperature combustion of HCCI engine resulted in low particulate and soot formations. Also, low specific fuel consumption was the most attractive of HCCI engines when compared with the conventional CI engines. Onishi et al. [5] conducted experiments to visualize the process of combustion on a conventional SI engine in both SI and HCCI modes using the schlieren photography method. They reported that very wee defined propagation of flame was found in SI operation mode but no visible propagation of flame was found in HCCI mode of combustion, proving volumetric combustion of HCCI engines [ 5, 6].

Swirl helps in homogeneous mixture formation of the fuel and air [ 7]. It also helps in NO_x emission reduction [ 8]. The increase in swirl ratio reduces the peak temperatures by increasing the heat transfer to the combustion chamber. This leads to a low temperature combustion process resulting in low NO_x emissions [ 9].

Performing these explorations (under different operating parameters with induction induced swirl) solely in the laboratory would be inefficient, expensive, and impractical since there are many variables that exhibit complex interaction. Therefore, a CFD tool Star-CD is chosen for the analysis. Several modifications were made to Star-CD es-ice module so that it could be used for HCCI engine modeling. The different combustion models which are well developed for predicting engine processes are transient interactive flamelets (TIF) model, digital analysis of reaction system-transient interactive flamelets model (DARS-TIF), G-equation model [ 10], extended coherent Flame combustion model-3 Zones [ 11] and the equilibrium-limited ECFM (ECFM-CLEH) [ 12, 13]. Each model has its own limitations and is suitable for a specific set of problems. Generally speaking, ECFM-3Z and ECFM-CLEH can be used for all types of combustion regime whereas ECFM-3Z is mostly suitable for homogeneous turbulent premixed combustion with spark ignition and compression ignition. The applicability of different combustion models is shown in Table 1. Owing to its wide applicability, in this paper, ECFM- 3Z is used to study the effect of swirl motion of intake charge on emissions and performance of the HCCI engine. Figure 1 depicts the schematic representation of the three zones of the ECFM- 3Z model. This model is capable of simulating the complex mechanisms like turbulent mixing, flame propagation, diffusion combustion and pollutant emission that characterize modern IC engines.

Induction induced swirl has a predominant effect on mixture formation and rapid spreading of the flame front in the conventional combustion process of a CI engine. This has been well documented in the literature. However, it is observed that no work has been done on the effect of swirl in HCCI mode.

The main objective of the present paper is to analyze the effect of induction induced swirl on controlling the peak pressures of the HCCI engine under varying operating parameters. Because of the volumetric combustion, the peak pressures inside the combustion chamber are very high. This is one of the limitations of the HCCI engine which reduces the engine life and demands robust engine design. In this regard, a computational attempt is made to control the peak pressures in terms of induction induced swirl along with other parameters.

Methodology

The software used in the present work makes use of computational fluid dynamics with the finite volume approach. The respective governing equations for conservation of mass, momentum, energy and species are solved consecutively in a solver, es-ICE, an expert system developed for internal combustion engines. The standard k-ω model is one, in which the turbulent Reynolds number forms the turbulent kinetic energy and turbulent dissipation rate equations are used in conjunction with the algebraic ‘law of the wall’ representation of flow, heat and mass transfer for the near wall region. The ECFM-3Z model is a general purpose combustion model capable of simulating the complex mechanisms of turbulent mixing, flame propagation, diffusion combustion and pollutant emission that characterize modern internal combustion engines. ‘3Z’ stands for three zones of mixing, namely the unmixed fuel zone, the mixed gases zone, and the unmixed air plus exhaust gas recirculation (EGR) zone. The three zones are too small to be resolved by the mesh and are, therefore, modeled as sub-grid quantities. The mixed zone is the result of turbulent and molecular mixing between gases in the other two zones and is the zone where combustion takes place. The flame propagation phase is modeled by the flame surface density transport equation incorporating the theoretical flame speed. The engine specifications considered for the analysis are listed in Table 2. The analysis is conducted from the second cycle after the engine has started.

CFD model set-up

The computational mesh consists of 128000 cells representing 1/6th of piston bowl created in STAR-CD by generating a spline based on the piston bowl shape. A 2D template is cut by the spline to cut the 3D mesh into 40 radial cells, 160 axial cells, 5 top dead center layers and 40 axial block cells. The piston bowl shape and 3D mesh of the piston bowl sector are illustrated in Fig. 2.

The energy efficiency of the engine is analyzed by gross indicated work per cycle (W) calculated from the cylinder pressure and piston displacement using Eq. (1).

(1) W={\displaystyle \frac{\mathrm{\pi }a{B}_{\mathrm{2}}}{\mathrm{8}}}{\displaystyle {\int }_{{\theta }_{\mathrm{1}}}^{{\theta }_{\mathrm{2}}}p(\theta )\left[\mathrm{2}\sin (\theta ){\displaystyle \frac{\mathit{a}\sin (\mathrm{2}\theta )}{\sqrt{l^{\mathrm{2}}-a^{\mathrm{2}}{\sin }^{\mathrm{2}}}(\theta )}}\right]}\mathrm{d}\theta ,

where a,l, and B are the crank radius, connecting rod length and cylinder bore, respectively, and q₁ and q₂ are the beginning and the end of the valve-closing period occurrences in terms of crank angles representing starting and end.

The indicated power per cylinder (P) is related to the indicated work per cycle by using Eq. (2).

(2) P={\displaystyle \frac{WN}{\mathrm{60000}{n}_R}},

where n_R=2 is the number of crank revolutions for each power stroke per cylinder and N is the engine speed (r/min). The indicated specific fuel consumption (ISFC) is expressed in Eq. (3).

(3) \mathrm{ISFC=}{\displaystyle \frac{\mathrm{30}{m}_{\mathrm{fuel}}N}{P}\mathrm{.}}

In Eq. (1), the power and ISFC analyses can be viewed as being only qualitative rather than quantitative in this paper.

Modeling strategy

The STAR-CD used in the present study has integrated several sub models such as turbulence, fuel spray and atomization, wall function, ignition, combustion, NO_x, and soot models for various types of combustion modes in CI as well as SI engine computations. As initial values of k and ϵ are not known, the turbulence initialization is done using the I-L model. For this purpose, local turbulence intensity, I, and length scale, L, are related as

(4) k_{\infty }=3I^2{V}_{\infty }^{\mathrm{2}}/2,

(5) {\epsilon }_{\infty }=C_{\mu }^{\mathrm{3}/\mathrm{4}}(k_{\infty }^{\mathrm{3}/\mathrm{2}}/L).

This practice will ensure that k, ϵ and the turbulent viscosity m_t, will all scale correctly with V_∞, which is desirable from both the physical realism and numerical stability point of view. Moreover, the turbulent intensity is defined using the same velocity vector magnitude as that of stagnation quantities.

The combustion is modeled using ECFM-3Z. As far as fluid properties are concerned, the ideal gas law and temperature dependent constant pressure specific heat (C_p) are chosen.

The ECFM-3Z incorporates the following models in its operation (Table 3).

Exhaust gas recirculation is mainly used to reduce the NO_x emissions and to improve the auto ignition. There are two EGR models as stated below.

(a) Variable composition—this model considers the components present in the EGR mixture. In this model up to six EGR scalars, namely EGR_O₂, EGR_CO₂, EGR_CO, EGR_H₂O, EGR_H₂ and EGR_N₂ are defined. These scalars are then solved by transport equations. The EGR_O₂ scalar does not take part in the reaction.

(b) Fixed composition—the EGR composition is defined by entering values for the mass fraction of each component in the O₂, CO₂, H₂O, N₂, CO and H₂ boxes. The sum of the supplied values must be equal to 1.0.

In the present analysis the variable composition model is used. In this model, EGR is considered as the mass of the re-circulating exhaust gas (m_egr) divided by the total mass that enters the cylinder (m_I).

Thus,

(6) \mathrm{egr}={\displaystyle \frac{m_{\mathrm{egr}}}{m_I}},

where

(7) m_I=m_{\mathrm{air}}+m_{\mathrm{egr}}+m_f\mathrm{.}

For individual species

(8) m_{\mathrm{egr}(f)}=\mathrm{egr}·m_{E(f)},

(9) m_{\mathrm{egr}({\mathrm{O}}_2)}=\mathrm{egr}·m_{E({\mathrm{O}}_2)},

(10) m_{\mathrm{egr}({\mathrm{C}\mathrm{O}}_2)}=\mathrm{egr}·m_{E({\mathrm{C}\mathrm{O}}_2)},

(11) m_{\mathrm{egr}({\mathrm{N}}_2)}=\mathrm{egr}·m_{E({\mathrm{N}}_2)},

(12) m_{\mathrm{egr}({\mathrm{H}}_2\mathrm{O})}=\mathrm{egr}·m_{E({\mathrm{H}}_2\mathrm{O})},

where m_f = mass of the fuel.

Subsequently, the mass fractions can be determined by dividing the mass of each individual exhaust gas by the total mass of all exhaust gases.

Initial and boundary conditions

To begin with, an absolute pressure of 1.02 bar, an EGR of 0%, a temperature of 353 K, and an equivalence ratio of 0.26 are taken as initial values. The fixed boundary wall temperatures are taken with combustion dome regions of 450 K, piston crown regions of 450 K, and cylinder wall regions of 400 K. The Angleberger wall function mode is considered [ 14]. The ‘two-layer’ and low Reynolds number approaches, with no-slip conditions are applied directly, whose boundary layers are computed by solving the mass, momentum and turbulence equations. The hybrid wall boundary condition which is a combination of two layered and low-Reynolds number wall boundary conditions is considered in this analysis. This hybrid wall boundary condition removes the burden of having to ensure a small enough near-wall value for y⁺ (by creating a sufficiently fine mesh next to the wall). The y⁺ independency of the hybrid wall condition is achieved either by using an asymptotic expression valid for 0.1<y⁺<100 or by blending low-Reynolds and high-Reynolds number expressions for shear stress, thermal energy and chemical species wall fluxes. This treatment provides valid boundary conditions for momentum, turbulence, energy and species variables for a wide range of near-wall mesh densities.

Standard wall functions are used to calculate the variables at the near wall cells and the corresponding quantities on the wall. The initial conditions are specified at IVC, consisting of a quiescent flow field at pressure and temperature for full load condition.

Validation of ECFM-3Z, compression ignition model

STAR-CD is a well-known commercial CFD package being adopted by many renowned researchers and well established research organizations in the field of automotive IC engines. The results obtained by using this package are validated with the experimental results by many authors like Pasupathy Venkateswaran and Nagarajan [ 8], Zellat et al. [ 15], Bakhshan and Tarahomi [ 16]. A comparison of the CI engine in HCCI is done in this paper considering the extended coherent flame combustion three zones, and compression model for combustion analysis. The present paper deals with the simulation of CI engine in HCCI mode, using a fuel vaporizer to achieve excellent HCCI combustion in a single cylinder air-cooled direct injection diesel engine. No modifications are made to the combustion system. Ganesh and Nagarajan [ 17] conducted experiments with diesel vapor induction without EGR and diesel vapor induction with an EGR of 0%, 10% and 20%. The validation of the present model with the experimental results of Ganesh and Nagarajan [ 17] was done by considering all the engine specifications.

Ganesh and Nagarajan [ 17] considered a vaporized diesel fuel with air to form a homogeneous mixture and inducted into the cylinder during the intake stroke. To control the early ignition of diesel vapor-air mixture, the cooled (30°C) EGR technique was adopted. For the validation purpose, the results are compared with respect to engine performance and emissions in Fig. 3. It is observed that the simulated results are in good agreement with the experimental ones. In Fig. 3, EDVI represents the experimental diesel vapor injection and SDVI represents simulated diesel vapor induction at respective EGR concentrations.

Results and discussion

In the present paper the effect of induced swirl in controlling peak pressures under the operating parameters like engine speed, equivalence ratio, exhaust gas recirculation, piston bowls are considered. The swirl ratios ranging from 1 to 4 are considered for the analysis. The simulation results of the ECFM-3Z model are discussed below.

Engine speed

The variation of peak pressures of the reentrant piston bowl with engine speed for swirl ratios of 1 to 4 are plotted in Fig. 4. From Table 4, it can be observed that lower engine speeds and lower swirl ratios are favorable in controlling the peak pressures. The peak pressures obtained at different engine speeds and swirl ratios are summarized in Table 4. The peak pressure decreases with the increase in engine speed irrespective of the swirl ratio, but the percentage increase in peak pressures is less at higher swirl ratios [ 18]. As the swirl ratio increases, the reduced peak pressures are obtained at all engine speeds, but the decrease in peak pressures is high at higher engine speeds [ 19]. The reason for this phenomenon is the increase in turbulence owing to increased wall heat transfer due to increased swirl intensity.

Exhaust gas recirculation (EGR)

The variation of peak pressures of the reentrant piston bowl with EGR concentration and with swirl ratios of 1 to 4 are depicted in Fig. 5 and the peak pressures obtained at different EGR concentrations and swirl ratios are listed in Table 5. From Table 5, it can be observed that higher EGR concentrations and higher swirl ratios are favorable in controlling the peak pressures.

The peak pressure decreases with the increase in EGR concentration irrespective of the swirl ratio, but percentage reduction in peak pressures are high at higher EGR levels with higher swirl ratios. This can be attributed to the release of reduced net energy with the decrease in volumetric efficiency of the engine at higher EGR levels and owing to oxygen availability for the combustion process. As swirl ratio increases, lower peak pressures are obtained at any EGR concentration, but percentage reduction in peak pressures is high at higher swirl ratios and higher EGR concentrations. The reason for this phenomenon is the increase in turbulence owing to increased wall heat transfer because of increased swirl ratios [ 20, 21]. A modest shift is also discovered in the occurrence of maximum peak pressure towards TDC with the increase in EGR concentrations [ 22].

Equivalence ratio (φ)

The variation of peak pressures of the reentrant piston bowl with equivalence ratio and with swirl ratios of 1 to 4 are demonstrated in Fig. 6 and the peak pressures (MPa) at different equivalence ratios and swirl ratios are listed in Table 6. From Table 6, it can be observed that lower equivalence ratios and higher swirl ratios are favorable in controlling the peak pressures. The peak pressure increases with the increase in equivalence ratio irrespective of the swirl ratio, but the percentage increase in peak pressures is marginal at all swirl ratios [ 23, 24]. The reason for this is that the increased equivalence ratio injects more fuel leading to a rich mixture formation process, and the combustion of more quantity of fuel leads to an increase in peak pressures.

As swirl ratio increases, the decrease in peak pressures are obtained at all equivalence ratios, but percentage reduction in peak pressures is high at high swirl ratios and lower equivalence ratios. This may again be caused by the increased turbulence owing to the increased wall heat transfer losses because of increased swirl ratios.

Piston bowl shape

The variation of peak pressures of the three different piston bowls with swirl ratios of 1 to 4 are displayed in Fig. 7 and the in-cylinder pressures at different piston bowls and swirl ratios are listed in Table 7. From Table 7, it can be observed that spherical piston bowl and higher swirl ratios are favorable in controlling the peak pressures. The spherical piston bowl is observed to be successful in controlling the peak pressures with the increase in swirl ratio [ 8, 25]. As swirl ratio increases, lower peak pressures are obtained for all the piston bowls, but the percentage decrease in peak pressures is higher for spherical piston bowl than reentrant and square piston bowls. The reason for this phenomenon is the increase in turbulence owing to increased wall heat transfer losses because of increased swirl ratios.

Conclusions

High peak pressures of HCCI engine due to volumetric combustion may damage the engine and thereby demands a rigid engine construction. Reduced engine life is one more problem because of high peak pressures. A computational analysis is undertaken to find the impact of induced swirl in controlling peak pressures of HCCI engine in this paper. The analysis of peak pressure control of an HCCI engine with induced swirl motion under different operating conditions has been done using the three zone extended coherent flame combustion model. The present study revealed that ECFM-3Z of STAR-CD was well suitable for HCCI mode of combustion with necessary modifications, in coherence with the existing literature. The swirl ratio had a considerable impact in controlling the peak pressures of HCCI engine. A reduction of approximately 21% in peak pressures was achieved when a swirl ratio of 4 with an EGR of 30% was adopted when compared to a swirl ratio of 1 with an EGR of 0%. The effect of four operating parameters, i.e., engine speed, exhaust gas recirculation, equivalence ratio and piston bowl geometry under varying swirl ratios suggested lower engine speeds, higher EGR concentrations, lower equivalence ratios, spherical piston bowl and higher swirl ratios are favorable in controlling the peak pressures. The adoption of high swirl ratios associated with high EGR levels would lead to significant reduction in peak pressures in HCCI mode.

References

[1] Karthikeya Sharma T, Amba Prasad Rao G, Madhu Murthy K. Performance of HCCI diesel engine under the influence of various working and geometrical parameters. International Journal on Theoretical and Applied Research in Mechanical Engineering, 2012, 1(1): 2319–3182

[2] Epping K, Aceves S, Bechtold R, Dec J E. The potential of HCCI combustion for high efficiency and low emissions. SAE Technical Paper, 2002, Paper No. 2002-01-1923

[3] Liu B, Su W, Wang H, Huang H. Characteristics and energy distribution of modulated multi-pulse injection modes based diesel HCCI combustion and their effects on engine thermal efficiency and emissions. Frontiers of Energy and Power Engineering in China, 2007, 1(4): 420–427

[4] Heywood J B. Internal Combustion Engine Fundamentals. New York: McGraw-Hill, Inc., 1988

[5] Onishi S, Jo S H, Shoda K, Jo P D, Kato S. Active thermo-atmosphere combustion (ATAC)—a new combustion process for internal combustion engines. SAE Paper, 1979, Paper No. 790501

[6] Karthikeya Sharma T, Amba Prasad Rao G, Madhu Murthy K. Combustion analysis of ethanol in an HCCI engine. Trends in Mechanical Engineering and Technology, 2013, 3(2): 6

[7] Manimaran R, Thundil Karuppa Raj R. Computational studies of swirl ratio and injection timing on atomization in a direct injection diesel engine. Frontiers in Heat and Mass Transfer, 2014, 5(1): 7

[8] Venkateswaran S P, Nagarajan G. Effects of the re-entrant bowl geometry on a DI turbocharged diesel engine performance and emissions—a CFD approach. Journal of Engineering for Gas Turbines and Power, 2010, 132(12): 122803

[9] Maroteaux F, Noel L, Ahmed A. Numerical investigations on methods to control the rate of heat release of HCCI combustion using reduced mechanism of n-heptane with a multidimensional CFD code. Combustion Theory and Modelling, 2007, 11(4): 501–525

[10] Duclos J P, Zolver M, Baritaud T. 3D modeling of combustion for DI-SI engines. Oil & Gas Science and Technology, 1999, 54(2): 259–264

[11] Colin O, Benkenida A. The 3-zone extended coherent flame model (ECFM3Z) for computing premixed/diffusion combustion. Oil & Gas Science and Technology – Rev. IFP, 2004, 59(6): 593–609

[12] Abouri D, Zellat M, Duranti S, Ravet F. Advances in combustion modeling in STAR-CD: validation of ECFM CLE-H model to engine analysis. In: 18th International Mutidimentional Engine User’s Meeting at the SAE Congress. Detroit, USA, 2008

[13] Subramanian G, Vervisch L, Ravet F. New developments in turbulent combustion modeling for engine design: ECFM-CLEH combustion sub-model. SAE Technical Paper, 2007, Paper No. 2007-01-0154

[14] Moureau V, Lartigue G, Sommerer Y, Angelberger C, Colin O, Poinsot T. Numerical methods for unsteady compressible multi-component reacting flows on fixed and moving grids. Journal of Computational Physics, 2005, 202(2): 710–736

[15] Zellat M, Duranti S, Liang Y, Kralj C, Schmidt G, Duclos J M. Towards a universal combustion model in STAR-CD for IC engines: from GDI to HCCI and application to DI Diesel combustion optimization. In: Proceedings of 14th SAE International Multidimensional Engine User’s Meeting. Detroit, USA, 2005

[16] Bakhshan Y, Tarahomi A R. Multi-dimensional simulation of n-heptane combustion under HCCI engine condition using detailed chemical kinetics. Journal of Engine Research, 2011, 22: 3–12

[17] Ganesh D, Nagarajan G. Homogeneous charge compression ignition (HCCI) combustion of diesel fuel with external mixture formation. Energy, 2010, 35(1): 148–157

[18] Viggiano A, Magi V. A comprehensive investigation on the emissions of ethanol HCCI engines. Applied Energy, 2012, 93: 277–287

[19] Sjöberg M, Dec J E. Effects of engine speed, fueling rate, and combustion phasing on the thermal stratification required to limit HCCI knocking intensity. SAE Technical Paper, 2005, Paper No. 2005-01-2125

[20] Gu Y, Mayor J R, Dahm W J A. Turbulence-augmented minimization of combustion time in mesoscale internal combustion engines. In: 44th AIAA Aerospace Sciences Meeting and Exhibit. Reno, USA, 2006, AIAA-2006-1350

[21] Chen G, Iida N, Huang Z. Numerical study of EGR effects on reducing the pressure rise rate of HCCI engine combustion. Frontiers of Energy and Power Engineering in China, 2010, 4(3): 376–385

[22] Murali K B, Mallikarjuna J M. Effect of engine speed on peak tumble flows in a motored internal combustion engine-an experimental investigation using particle image velocimetry. Journal of Applied Fluid Mechanics, 2011, 4(1): 1–14

[23] Martinez-Frias J, Aceves S M, Flowers D, Smith J R, Dibble R. Equivalence ratio-EGR control of HCCI engine operation and the potential for transition to spark-ignited operation. SAE Technical Paper, 2001, Paper No. 2001-01-3613

[24] Christensen M, Johansson B. Influence of mixture quality on homogeneous charge compression ignition. SAE Transactions. Journal of Fuels and Lubricants, 1998, 107: 1–13

[25] Jaichandar S, Annamalai K. Influences of re-entrant combustion chamber geometry on the performance of Pongamia biodiesel in a DI diesel engine. Energy, 2012, 44(1): 633–640

RIGHTS & PERMISSIONS

Higher Education Press and Springer-Verlag Berlin Heidelberg