Combustion Characteristics, Emissions and Heat Release Rate

Apr 2, 2009 - Combustion Characteristics, Emissions and Heat Release Rate Analysis of a Homogeneous Charge Compression Ignition Engine with ...
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Energy & Fuels 2009, 23, 2396–2404

Combustion Characteristics, Emissions and Heat Release Rate Analysis of a Homogeneous Charge Compression Ignition Engine with Exhaust Gas Recirculation Fuelled with Diesel Miguel Torres Garcı´a,* Francisco J. Jime´nez-Espadafor Aguilar, and Toma´s Sa´nchez Lencero Escuela Superior de Ingenieros de SeVilla. AVda. Camino de los Descubrimientos, s/n. 41092 SeVilla ReceiVed June 23, 2008. ReVised Manuscript ReceiVed March 13, 2009

This paper focuses on the study of heat release rate (HRR) and in-cylinder pressure on the homogeneous charge compression ignition (HCCI) process in a modified diesel engine (Deuzt Diter FL1 906). A zerodimensional thermodynamic model was used to calculate the HRR. The effects of inlet temperature, compression ratio, and exhaust gas recirculation (EGR) on the HRR and engine-out emissions have been investigated experimentally at speed range (1200-2400 RPM), where the start of injection has been fixed at 45° before top dead center (BTDC). The EGR, a well-known and widely used method to depress NOx emission in diesel engines and in HCCI combustion mode, is also used as a basic method to control the ignition timing and burning rate. The total combustion duration was prolonged with an increase of EGR, and the start of combustion was delayed. The HCCI combustion mode shows a HRR clearly different from the HRR correspondent to conventional diesel combustion as it is approached in Wiebe’s function of two modes.

1. Introduction Autoignition combustion of homogeneous air-diesel fuel mixtures (HCCI) is a technology with great potential in NOx and soot emissions reduction,1 while still providing high diesellike efficiency. It is based on the self-ignition of a homogeneous air-diesel fuel mixture without an external ignition source. To a degree, the HCCI combustion process is able to combine the best features of a SI engine using gasoline and a diesel engine using diesel fuel. Achieving acceptable HCCI combustion with diesel fuel can be difficult for two main reasons. First, elevated temperatures would be required before significant vaporization occurs, making it difficult to form a premixed homogeneous charge. Second, diesel fuel has significant cool-combustion chemistry, leading to rapid autoignition once compression temperatures exceed about 800 K.2 This can lead to overly advanced combustion phasing and/or require reduced compression ratios and low intake temperatures. Nevertheless, due to its many potential advantages, substantial efforts have been made to understand diesel-fuelled HCCI and to advance some approaches to commercial application. The rate of heat release in HCCI combustion mode is not controlled, neither by the rate of fuel injected as in DI engines, nor by finite turbulent flame propagation as in SI engines. Although there are a lot of models to predict HCCI autoignition, mostly based on chemical kinetics, these are difficult to calibrate, not to mention determine CFD codes. In addition, at present there is not an analytical function to model the HRR in HCCI combustion mode, which makes * To whom correspondence should be addressed. Phone: 003-495-4486111; fax: 003-495-448-7243; e-mail: [email protected] (1) Thring, R. H. “Homogeneous Charge Compression Ignition (HCCI) Engines,” SAE Paper 892068, 1989. (2) Kelly-Zion, P.; Dec, J. “A computational Study of the Effect of FuelType on Ignition Time in HCCI Engines”. 28th International Combustion Symposium 2000, 1, 1187–1194.

necessary the development of intensive experimental tests3 to explore a range of control strategies. Performing these explorations solely in the laboratory would be inefficient, expensive, and impractical since there are many variables that exhibit complex interactions. Fundamental tools, such as CFD codes with detailed chemistry, need to be applied in order to provide insight into the combustion process. Codes of this nature are very computationally intensive and usually require some simplifications to expedite the solution while attempting to maintain accuracy. All these models present a serious drawback: the huge computational load that precludes efficient and quick analysis of the HCCI process.4-6 This paper combines a modeling approach based on a thermodynamic zero-dimensional (0D) formulation in a single zone with a detailed experimental study. Previous studies showed that HCCI combustion, being intrinsically homogeneous, shows only a marginal dependence on the in-cylinder charge motion (turbulence, swirl, etc.). Once the combustion has started, the main effect of turbulence to be considered is through the increase of heat losses.7,8 The model (3) Haifeng, Liu; Mingfa, Yao; Bo, Zhang; Zunqing, Zheng. Effects of Inlet Pressure and Octane Numbers on Combustion and Emissions of a Homogeneous Charge Compression Ignition (HCCI) Engine, Energy and Fuels 2008, 22, 2207–2215. (4) Chen, Huang; Xingcai, Lu; Zhen, Huang. New Reduced Chemical Mechanism for Homogeneous Charge Combustion Ignition Combustion Investigation of Primary Reference Fuels. Energy Fuels 2008, 22 (2), 935– 944. (5) Jincai Zheng, Weiying Yang, David L. Miller and Nicholas P. Cernansky. “A Skeletal Chemical Kinetic Model for the HCCI Combustion Process,” SAE Paper. 2002-01-0423, 2002. (6) Lund, C. M. “A General Computer Program for Calculating TimeDependent Phenomena Involving One-Dimensional Hydrodynamics, Transport, and Detailed Chemical Kinetics LLNL Report, UCRL-52504, August 1987. (7) Christensen, M.; Johansson, B.; 2002. The effect of combustion chamber geometry on HCCI operation. SAE Technical Paper. 2002-010425, 2002. (8) Kong, S. S.; Reitz, R. D.; Christensen, M.; Johansson, B., 2003. Modeling the effects of geometry generated turbulence on HCCI engine combustion. SAE Technical Paper. 2003-01-1088, 2003.

10.1021/ef801010m CCC: $40.75  2009 American Chemical Society Published on Web 04/02/2009

HRR Analysis of a HCCI Diesel Engine with EGR Table 1. Data of EN590 Diesel Properties diesel fuel


density (1 atm and 15 °C) stoichiometric A/F ratio formula LHV viscosity (at 40 °C) cetane index

820-845 kg/m3 14.5 CnH1.8n 42.5 MJ/kg 2-4.5 mm2/s 46

allows us to evaluate the HRR and to analyze the influence of operating conditions in a modified diesel engine (FL1 906) running in HCCI combustion mode. In the past decade, a considerable amount of research effort has gone into the investigation of HCCI combustion diesel fuel, where the difficulty in achieving HCCI combustion without accessory methods stands out. Several potential control systems have been proposed to control HCCI combustion without changing the cetane number of diesel fuel: intake charge heating system,9 internal EGR,10 variable compression ratio (VCR),11 and variable valve timing (VVT)12,13 to change the effective compression ratio and/or the quantity of hot exhaust gases retained in the cylinder. The first objective of this paper was to investigate experimentally the combustion characteristic and HRR analysis obtained of a 0D model in a HCCI engine with EGR fuelled with diesel. The second objective was to show some effects of EGR and operating conditions on the emissions characteristic reached in the detailed experimental study. 2. Experimental Study The experimental part of this work was based on the tests done on the Deutz FL1 906 engine.14-16 The original diesel engine was modified to adapt it to HCCI combustion, the injection point was fixed and different operative conditions such as engine speed, air-fuel ratio and intake temperature were tested. The specifications of the engine are the following: cylinder bore 95 mm, stroke 100 mm, displacement 708 cm3, nominal compression ratio 19:1, and rated power of 13 kW at 3000 RPM. Commercial diesel fuel was used in every test, and its properties are shown in Table 1. Systematic tests were carried out, the angle of injection was fixed to 45° BTDC, and the temperature of intake air was changed from the atmospheric temperature of 18 to 50 °C in three intervals. The (9) QIAN, Zuo-qin; Lu¨, Xing-cai. Characteristics of HCCI engine operation for additive, EGR, and intake charge temperature while using iso-octane as a fuel. Journal of Zhejiang UniVersity SCIENCE A 2006, 7 (Suppl. II), 252–258. (10) Kathi Epping, Salvador Aceves, Richard Bechtold, Jonh Dec. The Potential of HCCI Combustion for High Efficiency and Low Emissions, SAE Paper 2002-01-1923, 2002. (11) Myung Yoon, Kim; Kim, Jee Won; Chang Sik, Lee; Je Hyung, Lee. Effect of Compression Ratio and Spray Injection Angle on HCCI Combustion in a Small DI Diesel Engine. Energy Fuels 2006, 20 (1), 69– 76. (12) Haifeng, Liu; Mingfa, Yao; Bo, Zhang; Zunqing, Zheng. Effects of Inlet Pressure and Octane Numbers on Combustion and Emissions of a Homogeneous Charge Compression Ignition (HCCI) Engine. Energy Fuels 2008, 22 (4), 2207–2215. (13) Hiraya, K.; Hasegaw, K.; Urushihara, T.; Kakuho, A.; Itoh, T. A study of gasoline fuelled compression ignition engine- a trail of operation region expansion, Proceedings of the JSAE ConVention (in Japanese) 2001, 98-01, 9–14. (14) Fernando Cruz, Perago´n; Fco Jime´nez-Espadafor, Aguilar. A Genetic Algorithm for determining Cylinder Pressure in Internal combustion engines. Energy and Fuels 2007, 21, 110–120. (15) Suzuki, H.; Koike, N.; and Odaka, M. “Combustion control method of homogeneous charge diesel engines,” SAE Technical Paper. 980509, 1998. (16) Torres Garcia, M.; Chacartegui Ramirez, R.; Jimenez-Espadafor Aguilar, F.; Sanchez, Lencero. Analysis of the Start of Combustion of a Diesel Fuel in a HCCI Process through an Integral Chemical Kinetic Model and Experimentation. Energy Fuels 2008, 22 (2), 987–995.

Energy & Fuels, Vol. 23, 2009 2397 Table 2. Measurement Accuracy calibration range

instruments fuel consumption (AVL PIERBUG PLU 401/121) in-cylinder pressure (Kistler 4045A2) in-cylinder pressure (Kistler 6061B) crankshaft encoder (Kistler 2613B)

relative error (%)


0.05-23 kg/h



0-2 bar



0-250 bar

(0.5 bar




load was changed from low load up to the attainable maximum load for the engine in HCCI combustion mode, this is, from 5 Nm up to 20 Nm with increments of torque of 5 Nm. Increasing engine load, without EGR, produces an advance on the start of combustion, an increase of the maximum pressure, and also an advance of the crankshaft angle where maximum pressure is reached. The maximum attainable load was established in order to ensure the mechanical integrity of the engine. Maximum pressure has been limited at TDC to 110 bar, which corresponds roughly to 20 Nm with the engine running without EGR at 1200 rpm, 19:1 compression ratio, and 18 °C of charge temperature. The range of engine speeds studied is from 1200 to 2400 RPM in intervals of 300 RPM.

3. Experimental Apparatus The engine has been tested at different engine speeds and loads on a test bed equipped with a 25 kW dynamometer with load cell for torque measurement. Cylinder pressure (see Table 2) was measured with a water-cooled piezoelectric sensor (Kistler 6061B), and the pressure at the intake pipe was measured with a piezorresistive sensor (Kistler 4045A2). This provides the absolute reference pressure for the whole engine cycle. Both sensors are coupled through a charge amplifier to the acquisition system, IOTECH Daq 3000 (16 bit 1 MHz), which allows triggering the system with a TTL square signal from a crank angle encoder (Kistler 2613B) with a resolution of 0.5°. For each engine speed and load the HRR was evaluated from the combustion pressure chamber and a 0D model, averaging for 250 consecutive cycles. The TDC was fitted through motored tests to different engine speeds. The recirculated gases (EGR) were cooled in order to maintain the intake gases below the ambient temperature. The external EGR rate was evaluated by measuring the EGR mass recirculated and the total mass inlet. The formula is as follows:

EGR(% mass) )

m ˙ EGR × 100 m ˙ EGR + m ˙ Air


where m ˙ EGR and m ˙ Air are measured with a hot wire flow meter. Fuel consumption is measured with volumetric flow equipment (AVL Pierburg PLU 401/121).

4. Model for the Evaluation of the HRR For this evaluation, a thermodynamic 0d model has been used to carry out the analysis and calculate the thermodynamic parameters. The combustion chamber is considered to be a perfect mixture reactor with variable volume, with even pressure distribution, temperature, and concentration of chemical species (see Figure 1). Heat losses have been calculated using an improved Woschni correlation.17 The equations and simplifications that govern the mathematical model are needed due to the complex interaction between physics and chemical phenomena during combustion. The energy balance was applied (17) Woschni, G. “Universally Applicable Equation for the Instantaneous Heat Transfer Coefficient in the Internal Combustion Engine”, SAE Paper 670931, 1967.


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Figure 1. Control volume for 0D analysis.

to the variable volume of the chamber to predict the temperature evolution in each time step. The first principle of thermodynamics applied to control volume formed by the combustion chamber is ref 18 and was applied between inlet valve close and exhaust valve open: dQT dW dE ) + + dθ dθ dθ

∑ j

m ˙j h ω j


where dQT/dθ (J/°) is the net total thermal energy flow including heat losses and fuel contribution, dW/dθ is the power (J/°), m ˙j is the mass flow of the j flow (kg/s), hj is the enthalpy (J/kg) of the flow in or out of the system, and E (J) is the total energy of the system, U + mgz + mc2/2. Of these terms, only U is taken into account due to the homogeneous characteristics of the system considered. EGR and gas flow thermodynamic properties, as enthalpy and internal energy, have been evaluated taking into account their dependence with temperature. Also, the change on gas mixture composition during combustion has been included. The procedure followed is the one exposed in ref 18. The heat release rate dQT/dθ (J/grade) is given by eq 3, dQf dQW dQT ) dθ dθ dθ


where Qf is the heat released by the fuel. The heat loss model is written as follows: dQW 1 ) Achc(Tg - TW) dθ ω


The heat transfer coefficient hc (W/m2 K) is based on the correlation formula given by Junseok Chang.19 QW is the wall heat loss, Ac is the area in contact with the gases, Tg (K) is the gas temperature every time (modeled), TW (K) is the cylinder wall temperature, and ω (rad/s) is the average engine speed. The global heat transfer coefficient can be written as: hc(t) ) Rscaling L(t)-0.2p(t)0.8T(t)-0.73V(t)0.8


A scaling factor Rscaling is used for tuning of the coefficient to match specific engine geometry. Combustion-induced gas (18) Heywood, J. B. “Internal Combustion Engine Fundamentals”, Ed. McGraw-Hill Book Company, Singapur (Singapur), 1988. (19) Junseok Chang, Orgun Gu¨ralp, Zoran Filipi, Dennis Assanis, TangWei Kuo, Paul Najt, and Rod Rask. “New heat Transfer Correlation for an HCCI engine derived from measurements of instantaneous surface heat flux”. Sae Paper 2004-01-2996, 2004.

Figure 2. Cumulative heat release profile with net heat release calculated from the 0D model, Φ ) 0.35, a fuel consumption of 0.015 g/cycle, 1800 RPM, 0% EGR, and intake temperature of 18 °C.

velocity is a function of the difference between motoring and firing pressure.20 V(t) ) C1Sjp +

C2 VdTr (p - pmotoring) 6 prVr


The main idea for using this equation is to keep the velocity constant during the unfiring period of the cycle and to then impose a steep velocity rise once combustion pressure departs from motoring pressure. The subscript r denotes a reference crank angle, such as intake valve closing time. Figure 2 shows the net cumulative heat release obtained from a fuel consumption of 0.015 g/cycle, 1800 RPM, and 0% EGR. Combustion efficiency comes from a measured exhaust gas composition, 93-96%, so the energy released by the fuel per cycle (LHV ) 42.5 MJ/kg) is between 593 and 612 J. Crevice losses are considered very low,19 about 2%. Parameter Rscaling from eq 5 has been tuned so that the cumulative heat released by fuel (from HRR), plus cumulative modeled heat losses, plus modeled cumulative crevices energy losses (2% of the energy content of the fuel flow injected per engine cycle), plus energy loss from measured combustion inefficiency are equal to the energy released by the fuel. This analysis was repeated over the whole range of operating conditions tested. The definition given by Lu¨ Xingcai has been adopted for defining the start of combustion from the results.21,22 In the experimental results there is no kinetic information, therefore the start of combustion is defined from the HRR obtained from the registered in-cylinder pressure combined with a 0D model of the combustion system; the combustion starts when the cumulative energy released is 5% of the total. The combustion duration has been established as the angle between the start of combustion and that corresponding to 10% of the magnitude of the peak of HRR on the falling side of the curve. Other authors hold that the end of combustion is when the total (20) Huber, K.; Woschni, G.; Zeilinger, K. Investigations on Heat Transfer in Internal Combustion Engines under Low load and motoring conditions. SAE Paper 905018, 1990. (21) Lu¨, Xingcai; Chen, Wei; Ji, Libin; Huang, Zhen. The effects of external exhaust Gas recirculation and Cetane number improver on the gasoline homogeneous charge compression ignition engines. Combust. Sci. Technol. 2006, 178, 1237–1249. (22) Shimazaki, N.; Akagawa, H.; and Tsujimura, K. “An experimental study of premixed lean diesel combustion.” SAE Technical Paper. 199901-0181, 1999.

HRR Analysis of a HCCI Diesel Engine with EGR

Figure 3. Combustion chamber pressure versus crank angle with a compression ratio of 15:1, intake temperature of 18 °C, engine speed of 2100 RPM, an initial Φ ) 0.48, and constant fuel consumption of 0.028 g/cycle for different EGR rates.

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Figure 4. Combustion chamber pressure versus crank angle with a compression ratio of 15:1, intake temperature of 18 °C, engine speed of 1800 RPM, an initial Φ ) 0.49, and fuel consumption of 0.028 g/cycle for different EGR rates.

released energy is 95% of the product of heating value (Hp) and fuel mass.23 5. Results and Discussion 5.1. Influence of EGR on the HRR and NOx, CO, and HC Emissions. The EGR, a very well-known method for NOx reduction in diesel engines, is also a method in HCCI combustion mode, when fuelled with commercial fuel, to improve engine power, mainly because of the increase of the ignition delay. The experimental tests with EGR have been carried out from an initial load condition without EGR and progressively increasing the percentage of EGR. In every load condition fuel consumption, engine speed, injection fuel angle, and intake temperature have been maintained constants. The percentage of EGR necessary to produce delay on the ignition depends on fuel consumption and on the engine speed. At constant RPM, it is observed that at low loads the percentage of EGR needed is bigger than at high loads, reaching up to 65 mass percent to low loads contrasted to 35 mass percent that is necessary at high loads. Also, the effect of EGR is similar to that produced by an increment of the engine speed: both agents join together to increase the ignition delay. The percentage of EGR needed fundamentally depends on the thermal capacity of the composition of the exhaust gases (quantity of CO2, H2O, and O2); at low loads the exhaust gases have a minimal proportion of CO2 and H2O and, therefore, less thermal capacity. In general, cold EGR plays a significant role on HCCI combustion characteristics. As shown in Figures 3 and 4, which correspond to two different tests, the maximum in-cylinder pressure decreases, the ignition delay increases, the peak value of HRR diminishes with the introduction of cold external EGR, and the combustion process is prolonged. The increase of EGR rate produces an increment of ignition delay. This diminishes the amount of fuel burned before TDC (BTDC), which reduces the maximum combustion pressure. Also the temperature of the gases are lower, which decelerates the combustion process. Although combustion is prolonged with any increment of the EGR rate, the effect of the ignition delay on the engine torque more than offset the penalty on thermodynamic efficiency, and although the combustion process separates from the constant volume combustion, the specific fuel consumption improves in (23) Lu¨, Xingcai; Hou, Yuchun; Ji, Libin; Zu, Linlin; Huang, Zhen. Heat Release Analysis on Combustion and Parametric Study on Emissions of HCCI Engines Fuelled with 2-Propanol/n-Heptane Blend Fuels, Enegy and Fuels 2006, 20, 1870–1878.

Figure 5. Combustion ignition angle and combustion duration as a function of EGR rate and fuel consumption with an intake temperature of 18 °C, 15:1 compression ratio, and 1500 RPM.

the end. In both Figures 3 and 4, any increment of EGR produces an increment of engine torque, due to a reduction of the fuel burned and, therefore, of the pressure BTDC up to a threshold level that the torque reduces. This is associated to lack of oxygen, that is, more than about 20% of EGR at 2100 RPM reduces the inlet air and so there is not enough oxygen to burn the fuel. Figure 5 shows combustion ignition angle and combustion duration versus EGR rate for different fuel consumption. In both images, two different zones can be clearly seen: from 0 to 25% EGR and from this to the end. In the first part, the combustion duration and the ignition angle change lightly and almost linearly with the EGR rate, but in the second zone the changes are more intense with EGR rate, which can be explained by the reduction


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Figure 6. Effect of cooled EGR on NOx emission with different fuel consumption and 1800 RPM.

in charge temperature and probably in the lack of air that decelerates the combustion rate. According to chemical kinetics of HCCI combustion,24-27 combustion characteristics are dominated by the low and middle temperature reaction mechanism. With an increase of cold EGR rate, the mixture temperature rising rate during the compression process decreases because of the increase of heat capacity. Also, the oxidation and decomposition reaction rate were depressed for the final products such as CO2, NO, and H2O. In Figure 6, NOx emissions reached in HCCI combustion mode are shown. The EGR rate in HCCI mode has been optimized to get the maximum engine torque showing a reduction close to 100% and never surpassing 40 ppm with an EGR rate higher than 20%. In HCCI combustion, the air-fuel mixture is uniformly distributed and combustion starts in areas locally hotter but always with a very low equivalence ratio. The charge is an even mixture of recirculated exhaust gases, air, and fuel, so the maximum temperature is low. Any increment of EGR raises the heat capacity of the mixture, lowering the maximum temperature and NOx emissions. Figure 6 also shows that NOx emissions on HCCI combustion mode are very low, stemming from the superdiluted fuel/air mixture. This can be explained by the results shown in Figure 7, in which the combustion temperatures calculated from incylinder pressure data and 0D model were plotted. Slight knock combustions will lead to higher NOx emissions, but higher NOx emissions can be depressed by an appropriate EGR rate because the in-cylinder gas temperature can be reduced by cold EGR (see Figure 7). Figure 7 shows that maximum gas temperature diminishes with any increase of EGR rate. As the combustion temperature is well below the critical temperature ≈1800 K for NOx formation through the thermal (or Zeldovich) mechanism,28,29 little NOx is produced during the HCCI combustion operation. Figures 8, 9, and 10 show the effects of EGR on exhaust emissions of CO, HC, and smoke in HCCI combustion mode. (24) Daeyup Lee, and Shinichi Goto. (Mechanical Engineering Laboratory MITI) “Chemical Kinetic Study of a Cetane Number Enhancing Additive for an LGP DI Diesel Engine,” SAE Paper. 2000-01-0193, 2000. (25) Lei, Shi; Yi, Cui; Kangyao, Deng; Haiyong, Peng; Yuanyuan, Chen. Study of low emission homogeneous charge compression ignition (HCCI) engine using combined internal and external exhaust gas recirculation (EGR). Energy 2006, 31, 2665–2676. (26) Curran, H. J.; Gaffuri, P.; Pitz, W. J.; Westbrook, C. K. “A ComprehensiVe Modeling Study of n-Heptane Oxidation” Combustion and Flame 1998, 114, 149–177. (27) Curran, H. J.; Gaffuri, P.; Pitz, W. J.; Westbrook, C. K. “A ComprehensiVe Modeling Study of iso-Octane Oxidation” Combustion and Flame 2008, 129, 253–280.

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Figure 7. Effect of cooled EGR on in-cylinder gas temperature with 0.014 g/cycle consumption fuel, 1800 RPM, initial Φ ) 0.27, and intake temperature of 18 °C.

Figure 8. Effect of cooled EGR on smoke; 1800 RPM, intake temperature of 18 °C, and 19:1 compression ratio.

Figure 9. Effect of cooled EGR on CO emissions; 1800 RPM, 19:1 compression ratio, and intake temperature of 18 °C.

There is still a certain amount of smoke, but it is relatively insignificant. With an increase of EGR rate, smoke in HCCI engine is unchanged, remaining at ultralows levels.30 That is to say, cooled external EGR can control HCCI combustion and

(28) Miller, J. A.; Bowman, C. T. Mechanism and modeling of nitrogen chemistry in combustion. Prog. Energy Combust. Sci. 1989, 15, 287. (29) Turns, Understanding NOx formation in nonpremixed flame: Experiments and modeling. Prog. Energy Combust. Sci., 21, 361, 1995. (30) Egnell, R. The influence of EGR on heat released rate and NO formation in DI diesel engine. Society of automotive Engineers Inc., Warrendale, PA, SAE Paper. 2000-01-1807, 2000.

HRR Analysis of a HCCI Diesel Engine with EGR

Figure 10. Effect of cooled EGR on HC emissions; engine speed 1800 RPM, 19:1 compression ratio, and intake temperature of 18 °C.

Figure 11. Combustion chamber pressure and the HHR in HCCI combustion mode vs crank angle for constant engine speed (1200 RPM), 19:1 compression ratio, and different fuel consumption and intake temperature.

NOx emissions without penalizing smoke emissions.31 With regard to CO and HC emissions, both increase with EGR rate. Although temperature and pressure time profiles of the gases at different locations throughout the cylinder are quite similar,32 temperature diminishes with any EGR rate increase. This precludes oxidation of unburned hydrocarbons emerging during expansion and exhaust from crevices in the combustion chamber and from any oil layers on the cylinder. 5.2. Analysis of Inlet Temperature on the HRR and Emissions. Figure 10 shows the combustion chamber pressure and the HRR in HCCI combustion mode versus crank angle for 1200 RPM speed and different fuel consumption and intake temperatures. From these tests the effects of air-fuel equivalence ratio and inlet temperature can be analyzed. It can be observed in Figure 10 that for the same inlet temperature every increase in fuel consumption also produces an increase on engine torque and an advance of the ignition angle. This behavior is due to the heating of the engine with the load that in the end increases the temperature of the inlet air. The same tendency is observed with the air intake temperature. The increase from ambient temperature (18 °C) to 50 °C is followed by an advance of the ignition angle. Also, it can be appreciated that as the torque rises, the inlet temperature has less influence on the start of combustion, probably due to the (31) Sanghoo, Kook; Seik, Park; Choongsik, Bae. Influence of Early Fuel Injection Timings on Premixing and Combustion in a Diesel Engine. Energy and Fuels 2008, 22, 331–337. (32) Shuji Kimura, et al. Ultra-clean Combustion Technology Combining a Low-Temperature and Premixed Combustion Concept for Meeting Future Emissions Standards. SAE Paper. 2001-01-0200, 2001.

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Figure 12. Combustion duration in crank angle versus inlet temperature for different fuel consumption and constant engine speed (1200 RPM).

huge heating of the whole intake system with the load. Another relevant observation of Figure 10 is that the peak value of the heat release rate also increases with fuel consumption. This is due to two different effects: • The advance of the start of combustion causes cylinder pressure to go up before TDC. This increases the compression work of the piston on the cylinder gases, and so any increment of engine torque needs a higher increment of fuel consumption. • The increase of engine temperature with fuel consumption accelerates combustion kinetics and diminishes combustion duration.33 Figure 10 shows an advance of start of combustion between 15 and 20° BTDC for the whole test. This produces a penalty on the SFC, thus the ambient temperature, 18 °C in the tests, is the best one in order to improve the efficiency of the engine. Also this temperature avoids warming up of the intake. The total combustion duration, in crank angle, versus inlet temperature and constant engine speed (RPM) is illustrated in Figure 12. This figure reinforces the last conclusion: a huge reduction of angular combustion duration with the inlet temperature. It can also be observed that the higher the fuel consumption, the higher the angular combustion duration, although when the inlet temperature reaches 50 °C, a severe reduction of combustion duration is produced. Another conclusion can be gathered from the figure: there is a higher sensitivity of combustion duration with the inlet temperature at low loads than at high loads. This is due to the heating effect of the load on the intake air, so that (∂k)/(∂Tinlet)|p is positive but tends to a constant, where k is the effective combustion rate constant. Figure 13 shows the combustion chamber pressure and the HRR in HCCI combustion mode versus crank angle for the same fuel consumption and different air intake temperature and engine speed. For any engine engine speed, the increment of intake temperature from 18 to 50 °C yields an ignition advance that diminishes the brake torque, due to the increase of combustion pressure before TDC. Otherwise it can be appreciated that for any inlet temperature, every increase in engine speed yields a delay of the ignition angle, an increment of combustion duration, and a diminution of the maximum of the HRR. Because the fuel consumption is constant in this analysis per each intake temperature, the amount of energy liberated per cycle does not change, although a (33) Jacques Lavy, Christian Angelberger, Philippe Guibert, Smalaki Mokhtari. Towards a Better Understanding of Controlled Auto-ignition (CAI) Combustion Process From 2 Stroker Engine Results Analyses, SAE Technical Paper. 2001-01-1859, 2001.


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Figure 13. Combustion chamber pressure and HHR in HCCI combustion mode versus crank angle for a constant fuel consumption of 0.012 g/cycle, 19:1 compression ratio for different fuel consumption, and different intake temperature and engine speed.

Figure 14. Effect of the intake temperature on NOx emissions; 19:1 compression ratio, for different EGR and 1800 RPM vs fuel consumption.

reduction of heat loss with any increment of engine speed,34,35 which in the end could produce a heating effect, can be predicted. These observations allow us to affirm that the increment of ignition delay, in CA, with engine speed is controlled by chemical kinetics of the combustion process. Ignition takes place according to pressure and temperature conditions from which the combustion rate is more-or-less locked for any engine speed, but for the same crank angle interval, any increment of engine speed reduces its duration and so finally the combustion angle increases and reduces the maximum of the HRR. Torque increment with engine speed is very low, even if there is a small reduction from 1600 to 2000 rpm at 18 °C air intake. One of the main reasons for this behavior is the increment of mechanical losses with engine speed that follows a quadratic law.18 Figure 14 shows NOx emissions versus fuel consumption at different intake temperature and EGR rate. It can be observed that when the EGR is 20% in mass, NOx emissions remain constant with fuel consumption; without EGR, NOx emissions increase with fuel consumption. The intake temperature has little effect on NOx emissions with a high EGR rate, but has a relevant effect when there is not EGR. One possible explanation of this tendency is the increase of charge temperature when combustion (34) Thomas Morel, Syed Wahiduzzaman, Dale R. Tree, David P. Dewitt. Effect of speed, Load, and Location on Heat Transfer in a Diesel Engine-Measurement and Predictions. SAE Paper. 870154, 1987. (35) Young, Liu; Reitz, R. D. Modelling of heat conduction within chamber walls for multidimensional internal combustion engine simulations. Int. J. Heat Mass Transfer 1998, 41 (6-7), 859–869.

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Figure 15. Effect of the intake temperature on HC emissions; 19:1 compression ratio for different EGR and 1800 RPM vs fuel consumption.

Figure 16. Effect of the intake temperature on CO emissions, for different EGR, 19:1 compression ratio, and 1800 RPM vs fuel consumption.

starts earlier. A shorter delay increases the amount of fuel burned BTDC, which raises charge temperature and, for this reason, NOx emissions. Figure 15 shows the effect of the intake temperature on HC emissions. In general, an increase of intake temperature helps reduce the HC emissions. This can be explained because any increase of intake temperature reduces the start of combustion angle, which diminishes combustion duration. For this reason, there is less time for hydrocarbon quenching in zones of low temperature. Also, it can be observed that for low fuel consumption, HC emissions are very sensible to a change of intake temperature and increase with fuel consumption up to an almost constant HC emissions level. Very similar behavior can also be observed in CO emissions, as shown in Figure 16. 5.3. Analysis of Compression Ratio on the HRR and NOx, HC, and CO Emissions. A reduction of the compression ratio from 19:1 to 15:1 has beneficial effects for the improvement of engine power in HCCI mode.36,37 It will produce a delay of the start of combustion angle and would thus reduce the quantity of EGR needed to center the combustion.38 Therefore, (36) Rakopoulos, C. D.; Antonopoulos, K. A.; Rakopoulos, D. C. Experimental heat release analysis and emissions of a HDSI diesel engine fuelled with ethanol-diesel fuel blends. Energy 2007, 32, 1791–1808. (37) Nebjosa Milovanovic, and Rui Chen. A Review of Experimental and Simulation Studies on Controlled Auto-Ignition Combustion (2001). SAE Paper 2001-01-1890, 2001. (38) Aceves, S. M.; Flowers, D. L.; Westbrook, C. K.; Smith, J. R.; Pitz, W. J.; Dibble, R. HCCI Combustion and Emissions,” SAE Paper. 2000-01-0327, 2000.

HRR Analysis of a HCCI Diesel Engine with EGR

Figure 17. Combustion chamber pressure vs crack angle with compression ratio of 19:1, engine speed of 1500 RPM, and fuel consumption 0.029 g/cycle with Φ ) 0.55 for different EGR rates.

Figure 18. Combustion chamber pressure vs crank angle with compression ratio of 15:1, engine speed of 1500 RPM, and fuel consumption of 0.026 g/cycle with Φ ) 0.51 for different EGR rate.

there would be more available oxygen that would allow increasing the mass of fuel and also the power. Besides the nominal 19:1, two compression ratios have been studied: 15:1 and 12:1. With a compression ratio of 15:1 it is necessary to heat the incoming air, but once the combustion starts it is possible to remove the warm-up. If the compression ratio diminishes to 12:1, the start of ignition is excessively delayed and the combustion turns erratic and unstable. Also, the incoming air must be heated continually, so this compression ratio has been rejected as useful. Figure 17 represents the effect of the EGR on the combustion pressure and HRR to 1500 RPM, with constant fuel consumption of 0.029 g/cycle and original compression ratio of 19:1. The maximum torque is 15 Nm and is obtained with 30% EGR mass, whereas with the compression ratio 15:1 (see Figure 18) with constant fuel consumption of 0.025 g/cycle is 18 Nm, and only 19% EGR mass is necessary. For the same engine torque and engine speed, the fuel consumption per cycle is lower with 15:1 than the 19:1 compression ratio, thus efficiency improves. The reduction of the compression ratio in HCCI combustion mode to 15:1 allows reaching maximum combustion pressures lower than the ones reached with the nominal compression ratio. Likewise, the SFC (a specific fuel consumption of 260 g/kWh) with 15:1 compression ratio is lower than 19:1 due to the delay of start of combustion; also, the maximum pressures are lower. It can be observed that the peak of the HRR and also the area under the HRR curve (associated with fuel consumption) are minor with 15:1 compression ratio. With the 19:1 compression ratio, the start of combustion cannot be delayed due to lack of

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Figure 19. Effect of cooled EGR on NOx emission with different compression ratio, intake temperature 18 °C, and 1800 RPM.

Figure 20. Effect of the compression ratio on HC emissions, intake temperature 18 °C, and 1800 RPM.

oxygen (dash-line in Figure 17 and 30% EGR), but with the 15:1 compression ratio it is possible to retard the start of combustion, without lack of oxygen, and thus obtain the maximum engine power. In this case, if the EGR ratio surpasses 18%, the maximum torque will diminish because of the excessive combustion duration and low combustion pressure. It can be concluded that there will be an optimum compression ratio between 19:1 and 15:1 that will allow obtaining the maximum torque without lack of oxygen, and thus thermodynamic efficiency will improve. Reducing the compression ratio from 19:1 to 15:1 has also beneficial effects on NOx emissions and almost null consequences on HC and CO emissions. Figure 19 shows that for low EGR rates and for the same fuel consumption, any increment of EGR rate diminishes NOx emissions, but always lowers emissions level for 15:1 in contrast to the 19:1 compression ratio. This can be explained by the advance of the start of combustion with 19:1, which increases the amount of fuel burned before TDC and therefore the charge temperature that promotes NOx formation. For high EGR rates, NOx emissions are almost null for both compression ratios. Figure 20 shows HC emissions and it can be observed that the EGR rate has an important influence on these emissions, although they are almost insensible to compression ratio. However, when the EGR ratio is 0%, a small but appreciable increase in HC emissions with 15:1 compression ratio can be observed. This can be explained by the advance of the start of combustion with 19:1 compression ratio, which increases the amount of fuel burned before TDC and therefore the charge


Energy & Fuels, Vol. 23, 2009

temperature. This diminishes the time for hydrocarbon oxidation, limiting the quenching near the linear and promoting HC oxidation. CO emissions show very similar effects with compression ratio and are not shown in separate figures to avoid repetition. 6. Conclusions Overall, this work shows a methodology able to establish some parameters for optimum engine operation in HCCI combustion mode. This has been done through the analysis of HRR, combustion chamber pressure, and exhaust gas emissions. The parameters studied have been engine compression ratio, intake temperature, and EGR rate. The main results are summarized as follows: (1) In general, cold EGR plays a significant role on HRR shape in HCCI combustion mode. The introduction of EGR delays the start of ignition, prolongs the combustion duration, and is able to increase engine efficiency. This last result is because, although maximum combustion pressure diminishes, more work is developed along the working stroke. (2) For diesel fuel with cetane index higher than 46, it should not be heated by intake flow. Any increase on charge temperature makes that combustion start earlier penalizes indicated work. To lessen this tendency, the EGR ratio could be increased. However, this reduces the air flow up to the point that all the fuel can not be burnt, mainly because of the lack of oxygen; this causes a high efficiency penalty. (3) NOx emissions, of HCCI combustion mode and optimized EGR, are near zero level. Soot emissions are almost null for any air-fuel equivalence ratio and EGR rate. (4) The air-fuel equivalence ratio has a direct effect upon HCCI ignition angle, advancing start of combustion, and reducing the combustion duration when the air-fuel equivalence ratio diminishes. The richest air-fuel equivalence ratios give the highest combustion pressure and produce the maximum HRR peaks. The first effect has a negative consequence on engine power because pressure increases during the compression stroke. (5) The HCCI combustion mode shows a HRR that corresponds clearly to a premixed combustion process. This suggests that an exponential heat release rate law can be useful to simulate combustion process in HCCI mode. (6) High HC and CO emissions on HCCI combustion mode could be decreased by increasing the intake temperature or the compression ratio. (7) The minimum NOx emissions and maximum power for constant fuel consumption are reached with the minimum allowable compression ratio compatible with engine operation without intake heating. Acknowledgment. The work in this paper is a part of item CTQ2007-68026-CO2-02/PPQ within the I+D+i national plan in the period 2007-2009 and has been backed by the Spanish

Garcı´a et al. Government (Ministry of Science and Education). The authors are grateful to the Ministry of Science and Education of Spain for their financial support.

Appendix m ˙ ) mass flow (kg/s) Sjp ) mean piston speed (m/s) A ) area (m2) Ac ) common area of heat transfer every time (m2) ATDC ) after top dead center BDC ) bottom dead center BTDC ) before top dead center CA ) crank angle C1, C2 ) constant adjusted depending on the engine type CD ) discharge coefficient CFD ) computational fluid dynamics CO2 ) carbon dioxide Cp ) constant pressure specific heat capacity (J/kg K) Cv ) constant volume specific heat capacity (J/kg K) DI ) direct injection E ) total energy (J) EGR ) exhaust gas recirculation (% mass) H2O ) water hc ) experimental heat coefficient (W/m2 K) HCCI ) homogeneous charge compression ignition HRR ) heat release rate IVC ) intake valve closing k ) effective combustion rate constant (cm3/mol s) LHV ) lower heat value (MJ/kg) m ) mass in combustion chamber (kg) P ) pressure (bar) Q ) heat (J) Qf ) heat released by the fuel (J) QT ) total thermal energy (J) QW ) wall loss heat (J) RPM ) revolutions per minute SFC ) specific fuel consumption SI ) spark ignition T ) temperature (K) TDC ) top dead center Tg ) gas temperature every time (K) TW ) cylinder wall temperature (K) U ) internal energy (J) V ) volume (m3) Vd ) displacement (m3) W ) power (W) Greek Letters γ ) (CP)/(CV) specific heats ratio F ) density (kg/m3) θ ) crankshaft angle (radians) Φ ) fuel-air equivalence ratio ω ) engine speed (rad/s) Rscaling ) scaling factor EF801010M