A new heat release rate (HRR) law for homogeneous charge compression ignition (HCCI) combustion mode

ACCEPTED MANUSCRIPT A new Heat release rate (HRR) law for Homogeneous Charge Compression Ignition (HCCI) combustion mode Miguel Torres García*; Francisco José Jiménez-Espadafor Aguilar; Tomás Sánchez Lencero; José Antonio Becerra Villanueva. Thermal Power Group, Department of Energy Engineering. University of Seville. Escuela Técnica Superior de Ingenieros de Sevilla. Camino de los Descubrimientos, s/n. 41092 Sevilla, España. *Corresponding author. Tel.:+34954486111; fax: +34954487243. E-mail: [email protected] (Miguel Torres García) Abstract Homogeneous charge compression ignition (HCCI) engines are drawing attracting attention as the next-generation's internal combustion engine, mainly because of its very low NOx and soot emissions and also for improvement in engine efficiency. Much research has been carried out in order to go deeper in this combustion process using multizone models or CFD codes. These simulation tools, although they can give a detailed view of the combustion process, are very time consuming and the results depend a lot on the initial conditions. A previous step to be considered in the simulation of the HCCI process is a heat release law evaluated from results of the experiment and a zero dimensional model. This paper focuses on the development of a new heat release rate (HRR) law that models the HCCI process when the combustion chamber is considered as a homogeneous volume. The parameters of this law have been adjusted through an optimization process that has allowed to fit the combustion chamber ACCEPTED MANUSCRIPT pressure. All the engine operative conditions from low to full load have been successfully simulated with this HRR law, with the maximum error in the estimation of combustion chamber pressure less than 2%. Keywords: HRR (heat release rate); EGR (exhaust gas recirculation); HCCI (homogeneous charge compression ignition), in-cylinder NOx. Nomenclature p SMean piston speed (m/s) AcCommon area of heat transfer a1, a2Shape factors for HRR ATDC After top dead center BDC Bottom dead center BTDC Before top dead center C1, C2Constants CFD Computational Fluid Dynamics CO2Carbon dioxide DI Direct injection EGR Exhaust Gas Recirculation H2O Water hcHeat transfer coefficient ACCEPTED MANUSCRIPT HCCI Homogeneous Charge Compression Ignition HRR Heat Release Rate K1, K2Terms of HRR law LHV Lower Heat Value m  Mass flow (kg/s) Mp, Mpp Shape factors P,p Pressure (bar) QHeat (J) QpHeat release in HCCI combustion mode (J) QWWall heat loss (J) RPM Revolutions per minute SI Spark ignition Ttemperature (K) TDC Top dead center TgGas temperature (K) TWCylinder wall temperature (K) VVolume (m3) VdDisplacement WPower (W) Greek letter θCrankshaft angle (radians) θpDuration of the energy release (radians) ACCEPTED MANUSCRIPT Фfuel-air equivalence ratio ωAngular speed (rad/s) scaling  Scaling factor 1. INTRODUCTION The auto-ignition combustion of homogeneous air-diesel fuel mixtures (HCCI) is a technology with great potential in NOxand soot emission reduction [1]. It is based on self ignition of a homogenous air-diesel fuel mixture without an external ignition source. At present, there is not an analytical function for modelling the heat release rate in HCCI combustion mode which hinders the combustion analysis and makes the development of intensive tests and CFD calculus necessary. Gas temperature is very important in the start of combustion and chemical kinetics. For this reason, heat transfer inside the combustion chamber has a significant role in HCCI combustion mode. Much research activity has been carried out on global heat transfer models, e.g. Annand[2], Woschnic[3] or Hohenberg[4], which have been elaborated in engine conditions that vary significantly from those in HCCI combustion. The dominant heat transfer mechanism in HCCI combustion mode is forced convection from hot gases to the combustion chamber walls. Due to very low soot formation and relatively low bulk gas temperature, the radiation effect is very small, the opposite being the case in the conventional diesel combustion mode. For this reason, any HRR analysis for HCCI combustion mode should be developed through a dedicated heat transfer model. In this paper, that of Junseok Chang has been used [5]. ACCEPTED MANUSCRIPT Regarding the rate of heat release in HCCI combustion mode, it is not controlled by the rate of fuel injection as in DI engines, nor by finite turbulent flame propagation as in SI engines. The absence of an ignition control mechanism has led researchers 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 interaction. Fundamental tools, such as computational fluid dynamics codes (CFD) 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. Fully coupled CFD/kinetic models have employed a number of different combustion methodologies. Kusaka [6] had no explicit modification of the reaction rate due to mixing; the only mixing that occurred during combustion was that normally calculated by the CFD code. Results with a detailed kinetic scheme showed a more rapid combustion event than the experiment. Because of the relatively coarse grid that was used, there is the possibility that the mass in colder zones was underestimated. Kong et al. (2001) [7] coupled the KIVA-3V CFD code with kinetics modified for turbulence effects and found a similar effect; i.e., when no turbulence effect was used, they predicted a higher energy release rate than was seen in the experiment. All these models present a severe drawback: the huge computational load that precludes efficient and quick analysis of the HCCI process. In this paper, a theory-experimental model is developed which allows evaluating the HRR in HCCI combustion mode in a modified diesel engine (DEUTZ-DITER FL1 906) from an analytical law of HRR. In all experimental studies carried out on HCCI combustion mode, it is observed that the ACCEPTED MANUSCRIPT HRR is quite different from the one in diesel combustion mode. The HRR of the HCCI process presents only a premixed combustion shape while that of the diesel mode shows premixed and diffusive appearance combustion. The knowledge of an analytical HRR law validated in a wide range of operating conditions in HCCI combustion mode will be a great advance in knowledge of HCCI combustion and could allow developing a predictive computational model for this combustion process. This paper develops a new heat release rate (HRR) law that models the HCCI process. The parameters of this law have been adjusted through an optimization process that has allowed to fit the combustion chamber pressure. The paper is organised as follows: Engine experimental setup, where the test rig and experimental procedure is described. Model for the evaluation of the HRR. Here the model developed for HRR from experimental measurement is explained. Analytical law for HRR in HCCI combustion mode, where the new HRR is presented. Numerical solution procedure. Results and conclusions. 2. ENGINE AND EXPERIMENTAL SETUP The experimental part of this work is based on the tests carried out on the modified DEUTZ FL1 906 engine. The original diesel engine was modified to adapt it to HCCI combustion. The main characteristics of the engine are shown in Table 1 and the experimental installation is represented in Figure 1. The modified engine systems were: ACCEPTED MANUSCRIPT 1º Injection System: the injection pump was extracted from the engine block, and its design modified to control the injection point. The gear of the crankshaft coupling to the injection pump was substituted by a transmission belt. 2º Exhaust gas recirculation system (EGR): an EGR system was designed which included an exhaust gases refrigeration system. With this system, the EGR by-passed mass fraction and the temperature of the recirculated gases can be controlled. 3º Measurement equipment: different sensors were located in the engine to monitor the pressure in the combustion chamber, the injection pressure, needle rising, engine speed, exhaust temperature, fuel mass flow, the intake system, pressure evolution, temperature and air mass flow. Figure 1 Start of combustion is highly dependent on intake temperature. To carry out a full analysis of the intake temperature effect in different operating conditions, a heating system for the intake air was designed (figure 1). Also, a heat exchanger cooler for the EGR gases was installed to control the recirculated gases temperature and so to control the temperature of the cylinder intake mixture. The installation incorporates stagnation boxes to eliminate the strong pulsating waves of the mono-cylinder engine. Table 1 Tests were done with a commercial diesel fuel compatible with European Standard EN590. The recirculated gases (EGR) were cooled in order to control the intake temperature that never surpassed the ambient temperature. The external EGR rate is ACCEPTED MANUSCRIPT evaluated measuring the EGR mass recirculated and the total mass inlet. The formula is as follows: 100)(%    AirEGR EGR mm m massEGR    (1) where EGR m  and Air m  are measured with a hot wire flow meter 3. MODEL FOR THE EVALUATION OF THE HRR In this work, a thermodynamic zero-dimensional model has been used to carry out the analysis and evaluation of the parameters of an analytical law of HRR, following an iterative optimization process. The combustion chamber is considered as a perfect mixture reactor with variable volume, with even pressure distribution, temperature and concentration of chemical species (see Figure 2) with heat losses. Equations and simplifications that govern the mathematical model are needed due to complex interaction between physics and chemical phenomenon during combustion and are described in the following equations [8]: Figure 2 Thermal state equation Combustion chamber volume First Principle of Thermodynamics[9] Heat losses Blow by has not been considered. The heat loss model is written as follows: ACCEPTED MANUSCRIPT    1 Wgcc WTThA d dQ  (2) The heat transfer coefficient hc(W/m2K) is based on the correlation formula given by Junseok Chang [5]. W Qis the wall heat loss, Acis the area in contact with the gases, Tg(K) is the gas temperature, Tw(K) is the cylinder wall temperature and ω(rad/s) is the average angular speed. The global heat transfer coefficient can be written as: 8.073.08.02.0 )()()()()( tvtTtptLth scalingc    (3) A scaling factor scaling  is used for tuning of the coefficient to match specific engine geometry. Combustion-induced gas velocity is a function of the difference between motoring and firing pressure [10].   motoring rr rd ppp Vp TV C SCtv  6 )( 2 1(4) 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. Figure 3 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 measured exhaust gas composition, 93-96%, so the energy released by the fuel per cycle (LHV=42.5 MJ/kg) is between 593-612J. Crevice losses are considered very low [5], about 2%. Parameter scaling  from equation (3) has been tuned so that the cumulative heat released by fuel plus heat losses, plus 2% from crevices, is equal to the energy ACCEPTED MANUSCRIPT From Table 3 and for engine conditions in Figure 9 it can be observed that every EGR increment also increases engine torque. The increase of torque is due to the diminution of combustion pressure during the compression stroke because of the delay in start of combustion, effect a. Because of this power increase at the same engine speed and fuel consumption, there is also an improvement in the engine's specific fuel consumption to weight of effect c that reduces thermodynamic efficiency [21]. Effect b produces a diminution of the combustion chamber temperature and so there will be a diminution of NOxemissions [22]. Figure 8 Figure 9 Table 3 quantifies the error between measurement and simulation built from the new HRR law. The difference between maximum pressure measured and predicted is less than 0.8%, the error of the area under cylinder pressure is less than 2% but is reduced to 1.4% when cylinder pressure is between inlet valve opening (IVO) and exhaust valve closing (EVC). Table 3 7. CONCLUSIONS In this work, an experimental and simulation study has been carried out to compare the performance of a new HRR law that defines a proportion of slower combustion for HCCI engine modelling. The new HRR law was implemented in an engine model to evaluate performance in comparison to the experimental data obtained in detailed tests. The study showed that by describing a proportion of slower combustion with the new HRR proposed, it was possible to achieve a very good match to experimental data. The new HRR law allows predicting the cylinder pressure curve perfectly with minimum error. As has been shown, the HRR law depends on four parameters that can be related ACCEPTED MANUSCRIPT to any load condition. Research is in progress on the development of a predictive model of the engine in HCCI combustion mode. ACKNOWLEDMENTS 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 supported by the Spanish Government (Ministry of Science and Education). The authors are grateful to the Ministry of Science and Education of Spain for their financial support of this work. REFERENCES 1. Thring. 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Y Daisho. D Kawako. H Susuki. and Y Goto. Emissions suppression mechanism of premixed diesel combustion with variable valve timing. International Journal of Engine Research. October 2007 Vol 8 No 5. ISSN 1468-0874. List figures 1. Experimental installation. 2. Control volume for zero-dimensional analysis 3. Cumulative heat release profile with net heat release calculated from the zerodimensional model, Ф=0.35, fuel consumption of 0.015 g/cycle, 1800 RPM, 0% EGR and intake temperature of 18ºC. 4. Combustion chamber pressure versus crank angle with a compression ratio of 15:1, intake temperature of 18ºC, angular speed of 2100 RPM, initial Ф=0.48 and constant fuel consumption of 0.028 g/cycle for different EGR rates. 5. Combustion ignition angle and combustion duration as a function of EGR rate and fuel consumption with an intake temperature of 18ºC, compression ratio of 15:1, angular speed of 2100 RPM 6. Scheme of the optimization procedure 7. Measured and predicted cylinder pressure traces with a 1.3e-5kg/cycle of fuel consumption, 1800 RPM. 8. Measured and predicted cylinder pressure and HRR simulated in HCCI combustion mode (angular velocity is 2100 RPM, constant fuel consumption 2.75e-5 kg/cycle, initial Ф=0.57 and different EGR rate). ACCEPTED MANUSCRIPT 9. Measured and predicted cylinder pressure and HRR simulated in HCCI combustion mode (angular velocity is 1500 RPM, constant fuel consumption 1.8e-5 kg/cycle, initial Ф=0.3 and different EGR rate). List tables 1. Deutz FL1 906 Engine characteristics. 2. The Best-fit parameters from new HRR law for different tests. 3. Accuracy of numeric model. Figure 1. Experimental installation. ACCEPTED MANUSCRIPT Figure 2. Control volume for zero-dimensional analysis Figure 3 Cumulative heat release profile with net heat release calculated from the zerodimensional model, Ф=0.35, fuel consumption of 0.015 g/cycle, 1800 RPM, 0% EGR and intake temperature of 18ºC. ACCEPTED MANUSCRIPT Figure 4 Combustion chamber pressure versus crank angle with a compression ratio of 15:1, intake temperature of 18ºC, angular speed of 2100 RPM, initial Ф=0.48 and constant fuel consumption of 0.028 g/cycle for different EGR rates. ACCEPTED MANUSCRIPT 0 5 10 15 20 25 0 5 10 15 20 25 30 35 EGR % mass Ignition angle (º BTDC) 0.017 g/cycle; Φinitial=0.40 0.01 g/cycle; Φinitial=0.25 0.02 g/cycle; Φinitial=0.43 0.028 g/cycle; Φinitial=0.48 0 2 4 6 8 10 12 0 5 10 15 20 25 30 35 EGR % mass Combustion duration (crank angle) 0.01 g/cycle; Φinitial=0.25 0.017 g/cycle; Φinitial=0.40 0.02 g/cycle; Φinitial=0.43 0.028 g/cycle; Φinitial=0.48 Figure 5 Combustion ignition angle and combustion duration as a function of EGR rate and fuel consumption with an intake temperature of 18ºC, compression ratio of 15:1, angular speed of 2100 RPM. ACCEPTED MANUSCRIPT Figure 6. Scheme of the optimization procedure. Figure 7. Measured and predicted cylinder pressure traces with a 1.3e-5kg/cycle of fuel