Energy Fuels 2009, 23, 4917–4930 Published on Web 08/24/2009
: DOI:10.1021/ef9004016
Numerical Study on Spray Combustion Processes in n-Heptane and Dimethyl Ether Fueled Diesel Engines Yongwook Yu, Sungmo Kang, Yongmo Kim,* and Kwan-Soo Lee Department of Mechanical Engineering, Hanyang University, Haengdang-Dong, Sungdong-Ku, Seoul 133-791, Korea Received May 4, 2009. Revised Manuscript Received July 24, 2009
The present study numerically investigates the distinctive differences of dimethyl ether (DME) spray combustion processes compared to conventional hydrocarbon liquid fuels in direct-injection diesel engines. To realistically represent the physical processes involved in the spray combustion, this study employs the hybrid breakup model, the stochastic droplet tracking model, collision model, high-pressure evaporation model, and transient flamelet model with detailed chemistry. The present representative interaction flamelet model with the vaporization effects on turbulent combustion has been validated against measurements for a Cummins N-14 diesel engine. The sequence of the comparative analysis has been made to understand the overall spray combustion characteristics of DME fuel as well as to identify the distinctive differences of DME combustion processes in direct-injection diesel engines, compared to nheptane liquid fuel. On the basis of numerical results, the detailed discussions are also made in terms of the spray dynamics, evaporation, turbulent mixing, autoignition, spray flame structure, turbulence-chemistry interaction, and pollutant (soot, NOx) formation in the n-heptane and DME fueled diesel engines.
ignition, and turbulent combustion processes of DME fuel especially in the DI diesel engine conditions. There have been experimental1-5 and analytical6 studies to understand the characteristics of the DME spray combustions as well as to optimally design the DME injection and combustion systems. The numerical modelings for the DME spray combustion processes are relatively rare due to the lack of reliable and fully informative experimental data as well as the shortcomings of the combustion model to realistically simulate the DME spray combustion processes. Golovitchev et al.7,8 performed numerical simulations of DME spray combustion. The predictive capability of their spray combustion model was validated against experimental data in terms of liquid and vapor penetration and ignition in a constantvolume chamber. Kim et al.9 numerically and experimentally investigated the characteristics of the turbulent combustion processes of DME sprays. Numerical simulation of the spray development and ignition process of DME sprays was performed using a transient flamelet model together with the lowpressure vaporization model and the reduced chemical kinetic mechanism. The numerical results agreed reasonably well with the experimental data. However, there is still a lot of room to improve the physical submodels to realistically predict the physically complex DME spray combustion processes. Very recently, in order to identify the inherent,
1. Introduction Among oxygenated fuels, the simplest ether fuel, dimethyl ether (DME), has been attracting much attention as a clean alternative fuel for diesel engines. The cetane number of DME is high enough to operate conventional compression-ignition engines. The thermal efficiency of a DME-powered diesel engine is comparable to that of diesel fuel operation, and sootfree combustion can be achieved without any extra modifications. However, since DME has distinctly different spray combustion characteristics from conventional hydrocarbon liquid diesel fuels in terms of evaporation, ignition, vapor pressure, cetane number, oxygenate ingredient, heat release rate, and liquid density, the application of DME in diesel engines creates many problems associated with the fuel-air mixing processes. Although DME burns well in the combustion systems of direct-injection (DI) diesel engines at light and medium loads and all speeds, the combustion efficiency of DME fueled diesel engines with insufficient mixing could be deteriorated at high loads and high speeds. In this respect, more research is needed for spray dynamics, vaporization, *Corresponding author. E-mail:
[email protected]. Phone: þ82-2-2220-0428. Fax: þ82-2-2297-3432. (1) Arcoumanis, C. Alternative Fuels for Transportation. The second European Auto-Oil programme (AOLII); European Commission, 2000, Vol. 2. (2) Sorenson, S. C.; Glensvig, M.; Abata, D. Di-methyl Ether in the Diesel Fuel Injection Systems; SAE paper 981159, Society of Automotive Engineers: Warrendale, PA, 1998. (3) Wakai, K.; Nishida, K.; Yoshizaki, T.; Hiroyasu, H. Spray and Ignition Characteristics of Di-methyl Ether Injected by a DI Diesel Injector. Proceedings of the Fourth International Symposium COMODIA 98 , 1998; pp 537-542. (4) Kajitani, S.; Chen, Z.; Oguma, M.; Konno, M. A study of low compression-ratio di-methyl ether diesel engines. Int. J. Engine Res. 2002, 2, 1–11. (5) Wakai, K.; Nishida, K.; Yoshizaki, T.; Hiroyasu, H. Ignition Delays of DME and Diesel Fuel Sprays Injected by a D.I. Diesel Injector; SAE paper 1999-01-3600, Society of Automotive Engineers: Warrendale, PA, 1999. r 2009 American Chemical Society
(6) Teng, H.; McCandless, J. C.; Schneyer, J. B. Thermo-chemical Characteristics of Di-methyl Ether - An Alternative Fuel for CompressionIgnition Engines; SAE paper 2001-01-0154, Society of Automotive Engineers: Warrendale, PA, 2001. (7) Golovitchev, V. I.; Nordin, N.; Chomiak, J. Neat di-methyl ether: is it really diesel fuel of promise?; SAE Paper 982537, Society of Automotive Engineers: Warrendale, PA, 1998. (8) Golovitchev, V. I.; Nordin, N.; Chomiak, J.; Nishida, N.; Wakai, K. Evaluation of ignition quality of DME at diesel engine conditions. Proceedings of the Fourth International Conference of Internal Combustion Engines 99 (ICE99): Experiments and Modeling, 1999; pp 299-306. (9) Kim, Y.; Lim, J.; Min, K. A Study of the Dimethyl Ether Spray Characteristics and Ignition Delay. Int. J. Engine Res. 2007, 8, 337–346.
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intrinsic, and distinctive differences of DME spray combustion processes compared to conventional hydrocarbon liquid fuel in the diesel-like engine conditions, Yu et al.10 have made the comparative analysis for the combustion processes and flame structure of DME and n-heptane fuel spray jets in a constant volume chamber. However, the previous studies still underexplored the precise differences between DME and conventional diesel fuels in the real diesel engine conditions. In this regard, the present study numerically investigates the distinctive differences of DME spray combustion processes compared to conventional hydrocarbon liquid fuels in the direct-injection diesel engines. In order to evaluate the predicative capability for the spray dynamics, vaporization, autoignition, and combustion process in the DI diesel engine, the present RIF (representative intercation flamelet) model with the vaporization effects on turbulent combustion has been validated against measurements11 for a Cummins N-14 engine. Moreover, in order to understand the overall spray combustion characteristics of DME fuel as well as to identify the distinctive differences of DME combustion processes compared to the conventional hydrocarbon liquid fuels, the sequence of the comparative analysis has been systematically made for the n-heptane and DME spray combustion processes in the direct injection diesel engines. On the basis of numerical results, the detailed discussions are also made in terms of the spray dynamics, evaporation, turbulent mixing, chemical kinetics, autoignition, spray flame structure, turbulencechemistry interaction, and pollutant (soot, NOx) formation in the n-heptane and DME fueled diesel engines.
zation model, standard k-ε turbulent model, transient flamelet model,15,16 the detailed chemistry model, and the soot and NOx model. The detailed numerical procedure for the RIF approach,15,17,18 the high-pressure vaporization model,16 and comprehensive spray combustion model13 can be found elsewhere. All these physical models for the spray dynamics are implemented in the KIVA II code.13 Besides these state-ofthe-art spray dynamics, the high-pressure vaporization model has been adopted to correctly account for high-pressure vaporization processes in the context of the comprehensive spray combustion modeling. In order to realistically simulate the turbulence-chemistry interaction in the spray combustion processes, the RIF (representative interactive flamelet) model15 has been also employed. To account for the high-pressure vaporization processes in context with the comprehensive spray combustion modeling, the high-pressure vaporization model is utilized in this study. To calculate the heat and mass flux between a droplet and a gas field, a film correction presented by Abramzon and Sirignano19 is chosen. In the liquid vaporization model,20 it is important to calculate the physical properties accurately at both the vapor and liquid phases of each species. Internal circulation arising from shear force must be considered when a relative velocity exists between the droplet surface and a nearby gas. In order to include the internal circulation effect, the effective conductivity model is introduced by Abramzon and Sirignano.19 In the present study, the properties of each species at both vapor and liquid phase are calculated as a function of temperature and pressure. The appropriate mixing rules are also used for calculations of mixture properties.21 Thermodynamic equilibrium at the droplet surface requires that the fugacities of each species in the gas phase are equal to their fugacities in the liquid phase. Deviation between the latent heat of vaporization for pure components and the enthalpy of vaporization of a gas mixture is determined by employing the Peng-Robinson equation of state (EOS). In the soot modeling in conjunction with the flamelet approach, the rates of soot volume fraction and number density are expressed as a function of the mixture fraction and the scalar dissipation rate. The present study adopts the two-equation model which is represented by the Favre averaged transport equation of soot number density and soot volume fraction. The semiempirical soot model proposed by Moss et al.22 is employed to calculate the source terms for
2. Physical and Numerical Models The spray combustion involves the complex physical processes such as the atomization of the liquid fuel, droplet breakup, droplet dispersion by turbulence, droplet collision, evaporation, turbulent mixing, autoignition, and turbulence-chemistry interaction. In this study, all submodels for these important physical processes are implemented in the multidimensional Eulerian-Lagrangian formulation. The gas-phase equation is written in an Eulerian coordinate whereas the liquid phase is presented in Lagrangian coordinates. The two-way coupling between the two phases is described by the interphase source terms that represent the rate of momentum, mass, and heat transfer. The physical models used in the present study include the hybrid droplet breakup model,12 stochastic droplet tracking technique,13 O’Rourke’s droplet collision model,14 high-pressure vapori-
(16) Yu, Y.; Kim, S. K.; Kim, Y. Numerical Modeling for Autoignition and Combustion Processes of Fuel Spays in High-Pressure Environment. Combust. Sci. Technol. 2001, 168, 85–112. (17) Barths, H.; Antoni, C.; Peters, N. Three-Dimensional Simulation of Pollutant Formation in a DI-Diesel Engines Using Multiple Interactive Flamelets; SAE paper 982456, Society of Automotive Engineers: Warrendale, PA, 1998. (18) Kim, S. K.; Yu, Y.; Ahn, J. H.; Kim, Y. Numerical investigation of the autoignition of turbulent gaseous jets in a high-pressure environment using the multiple-RIF model. Fuel 2004, 83, 375–386. (19) Abramzon, B.; Sirignano, W. A. Approximate Theory of a Single Droplet Vaporization in a Convective Field: Effects of Variable Properties, Stefan Flow and Transient Liquid Heating. Proceedings of Second ASME-JSME Thermal Engineering Joint Conference,1987; Vol. 1, pp 11-18. (20) Kneer, R.; Schneider, M.; Noll, B.; Witting, S. Diffusion Controlled Evaporation of a Multicomponent Droplet: Theoretical Studies on the Importance of Variable Liquid Properties. Int. J. Heat Mass Transfer 1993, 36 (9), 2403–2415. (21) Reid, R. C.; Prausnitz, J. M.; Poling, B. E. The Properties of Gases & Liquids, fourth ed.; McGraw-Hill: New York, 1987. (22) Moss, J. B.; Stewart, C.; Syed, K. Flow field modeling of Soot formation at elevated pressure. 22nd Symposium (International) on Combustion, The Combustion Institute, 1988; pp 413-423.
(10) Yu, Y.; Kang, S.; Kim, Y.; Lee, K. S. Numerical Study on the Characteristics of Vaporization, Ignition, and Turbulent Combustion Processes in Dimethyl Ether (DME)-Fueled Engine Conditions. Energy Fuels 2008, 22, 3649–3660. (11) Venkatesan, C. P.; Abraham, J. An Investigation of the Dependence of NO and Soot Emissions from a Diesel Engine on Heat Release Rate Characteristics - I; , SAE Paper 2000-01-0509, Society of Automotive Engineers: Warrendale, PA, 2000. (12) Patterson, M. A.; Reitz, R. D. Modeling of the effects of fuel spray characteristics on diesel engine combustion and emission; SAE Paper 980131, Society of Automotive Engineers: Warrendale, PA, 1998. (13) Amsden, A. A.; O’Rourke, P. J.; Butler, T. D. KIVA II: A Computer Program for chemically Reactive Flows with Sprays; LA11560-MS, Los Alamos National Labs: Los Almos, NM, 1989. (14) O’Rourke, P. J. Collective Drop Effects on Vaporing Liquid Sprays; LA-9069-T, Los Alamos National Laboratory: Los Almos, NM, 1981. (15) Pitsch, H.; Barths, H.; Peters, N. Three-Dimensional Modeling of NOx and Soot Formation in DI-Diesel Engines Using Detailed Chemistry Based on the Interactive Flamelet Approach; SAE Paper 962057, Society of Automotive Engineers: Warrendale, PA, 1996.
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Table 1. Specification of Cummins N-14 DI Diesel Engine engine type bore (mm) stroke (mm) compression ratio displacement volume (cm3) injection nozzles injector hole diameter (cm) spray inclination angle (deg)
Cummins N-14 13.97 15.24 16.5 2340 8 0.02 76
Table 2. Experimental Conditions of Cummins N-14 DI Diesel Engine
A B C
RPM
Pi (MPa)
Ti (K)
mass of fuel (g)
SOI (°CA)
EOI (°CA)
1200 1200 1200
0.1244 0.1244 0.1244
408 408 408
0.0657 0.0657 0.0657
2 7 9.5
10 15 17.5
Figure 1. Computational domain of Cummins N-14 DI diesel engine at top dead center.
nucleation, coagulation, surface growth, and oxidation in these soot property transport equations. In this study, the n-heptane chemistry is represented by the detailed chemical mechanism23 comprising 114 elementary reactions with 43 species and 101 irreversible reactions with 13 additional species to predict NOx formation. The reaction scheme includes fuel oxidation, low temperature degenerate chain branching to describe autoignition, and a reaction mechanism for NOx formation, including thermal, prompt, nitrous NOx, and reburn by hydrocarbon radicals. In terms of the DME chemistry, the chemical mechanism24 comprises 336 chemical reactions with 78 chemical species, and the extended Zel’dovich mechanism25 is employed to predict NOx formation. In the present study, the RIF model15 has been employed to realistically simulate the turbulence-chemistry interaction in the spray combustion processes. Moreover, the Eulerian particle flamelet model (EPRM)17 utilizing the multiple RIFs model is to handle the spatial inhomogeneity of the scalar dissipation rate. In order to account for the vaporization effects on the turbulent spray combustion, the present study adopts the model proposed by Demoulin and Borghi.26 The mean species mass fractions are calculated by integrating the flamelet solution weighted with a presumed probability density function. The calculation procedure of the RIF model is
Figure 2. Cylinder pressure histories calculated by single and multiple RIF models for three different injection timings in a Cummins N-14 DI diesel engine.
performed interactively with the computational fluid dynamics (CFD) solver. During one time step of main CFD code, the flamelet equations are solved by the stiff ordinary differential equation (ODE) solver, in which the time step is subdivided adaptively into subcycles to resolve the much smaller chemical time scales. The detailed numerical procedure for the RIF approach15,17,18 and comprehensive spray combustion model10,21 including the RIF approach can be found elsewhere. In this work, in order to numerically investigate the transient turbulent reacting flows, the multiple RIF model is implemented in the KIVA code.13
(23) Liu, S.; Hewson, J. C.; Chen, J. H.; Pitch, H. Effects of Strain Rate on High-Pressure Nonpremixed n-Heptane Autoignition in Counterflow. Combust. Flame 2004, 137, 320–339. (24) https://www-pls.llnl.gov/?url=science_and_technology-chemistry-combustion. (25) Miller, J. A.; Bowman, C. T. Mechanism and Modeling of Nitrogen Chemistry in Combustion. Prog. Energy Combust. Sci. 1989, 15, 287–338. (26) Demoulin, F. X.; Borghi, R. Assumed pdf modeling of turbulent spray combustion. Combust. Sci. Technol. 2000, 158, 249–272.
3. Results and Discussion Validation Case I: n-Heptane Fueled Diesel Engines. Before we perform the parametric numerical study, to evaluate the predicative capability for the spray dynamics, vaporization, 4919
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75 84.5 17.8 373.3 6 0.128 78
Table 4. Experimental Conditions for t of the DME Fueled Engine engine speed (rpm) injection pressure (MPa) injection mass (mg/cycle) start of energizing
1500 35, 50 4, 8, 12 6°, 4°, 2° BTDC
Figure 3. Temporal evolutions of the scalar dissipation rates and the maximum temperatures of flamelets for 9.5° CA injection timing in a Cummins N-14 DI diesel engine.
Figure 5. Computational domain of the DME fueled diesel at top dead center.
Figure 6. Measured and predicted cylinder pressure histories in the DME fueled engine (Pinj = 35 MPa, injected mass = 8 mg, SOE = 4° BTDC).
has been validated against measurements11 for a Cummins N-14 engine. The engine specification and computational conditions are given in Tables 1 and 2. Computation starts at bottom dead center (BDC) and continues until the exhaust valve opens at 130° crank angle (CA). In experiments, a diesel fuel is used. Since the detailed chemical kinetics of diesel fuel are still uncertain, n-heptane is used as a fuel in numerical simulation. The orifice diameter is 200 μm, and the spray angle is 14°. Computations are performed for a 45° sector of the combustion chamber with the periodic boundary condition in the azimuthal direction, and the spatial grid system at top dead center (TDC) is presented in Figure 1.
Figure 4. Instantaneous contours of mean temperature for 7° CA injection timing in a Cummins N-14 DI diesel engine.
autoignition, combustion process, and pollutant (NOx and soot) formation in the DI diesel engine, the present RIF model with the vaporization effects on turbulent combustion 4920
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Figure 7. Instantaneous contours of mean temperature predicted by the multiple RIF model (Pinj = 35 MPa, injected mass = 8 mg, SOE = 4° BTDC).
Figure 9. Temporal evolutions of the scalar dissipation rates (solid line) and the maximum temperatures (dashed line) of DME and n-heptane fueled diesel engines for three different injection timings (0 2.0°, 4 7.0°, and O 9.5° ATDC).
To investigate the spatial inhomogeneities of the scalar dissipation rate on the spray combustion processes of the DI diesel engines, the present study utilizes two turbulent combustion models including the single RIF model and the multiple RIF model. In Figure 2a, the simulated cylinder pressure histories obtained by the single RIF model are compared with the measured data for three injection timings. The computed and measured pressure histories and ignition delay times have reasonably good agreement for all cases. However, the overestimations in the ignition delay time and cylinder pressure are quite noticeable for the relatively late injection timings (7° and 9.5° CA). These discrepancies could be mainly related to the use of the n-heptane fuel to simulate the multicomponent diesel fuel which has the different characteristics in vaporization, chemical kinetics, spray penetration, and two-phase air-fuel mixing processes. The neglect of the spatial inhomogeneity of scalar dissipation rate in the transient spray flame field is partly responsible for the overestimated ignition delay time and cylinder pressure. However, in terms of the cylinder pressure histories, the
Figure 8. Cylinder mean pressure and temperature histories of DME and n-heptane fueled diesel engines for three different injection timings.
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Figure 10. Instantaneous contours of mean temperature and droplet trajectory for 2° ATDC injection timing in DME and n-heptane fueled diesel engines.
present model noticeably overestimates the cylinder pressure after autoignition at the relatively late injection timing (7° CA). In conjunction with the single RIF model, neglect of the spatial inhomogeneity of scalar dissipation rate could be partially responsible for the overprediction of cylinder pressure. In order to improve the predicted ignition delay time and cylinder pressure histories, the multiple RIF model including vaporization effects on the turbulent combustion process is used to simulate the spray combustion processes for three injection timings. This multiple RIF approach adopts the Eulerian particle flamelet model utilizing 10 RIFs which is capable of accounting for spatial inhomogeneity of the scalar dissipation rate. Figure 2b displays the simulated cylinder pressure histories obtained by this multiple RIF model. The multiple RIF model predicts slightly earlier autoignition and slightly lower cylinder pressure than the single RIF model. When the spatial inhomogeneity of scalar dissipation rate is included through the multiple RIF procedure, the predicted ignition delay time and cylinder pressure history have noticeably better conformity with experimental data, compared to the single RIF approach. However, there are still large discrepancies between prediction and measurement which could be mainly caused by the use of the n-heptane fuel to simulate the multicomponent diesel fuel.
Figure 3 shows the temporal evolutions of the scalar dissipation rate and the maximum temperatures of flamelet for 9.5° CA injection timing predicted by the single RIF model and the multiple RIF approach utilizing 10 flamelets. At the initial stage of liquid fuel injection, the scalar dissipation rate rapidly increases and reaches to a peak value due to the high vaporization rate and the intense turbulence generated by the high-velocity injection process. The high injection velocity enhances the liquid jet disintegration, droplet breakup, interphase convective heat transfer, droplet vaporization, and gas-phase turbulence. Particularly, the droplet vaporization rate is greatly increased by reducing the vaporization characteristic time in the strong droplet breakup process as well as by increasing the interphase convective heat transfer. As a result, this high injection velocity abruptly increases the scalar dissipation rate by elevating these nonequilibrium effects such as the enhanced turbulence, the high spray evaporation rate and the large gradient of mixture fraction. Shortly after this initial injection period, the scalar dissipation rate gradually decreases to the slow-varying low value since the fuel vapor spreads out the ambient flow field and the turbulence is continuously dissipated. At the small scalar dissipation rates, the diffusive losses become so small that heat and radicals produced by the chemical reaction are continuously increased and the mixture field can be finally 4922
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Figure 11. Instantaneous contours of OH radical mass fraction for 2° ATDC injection timing in DME and n-heptane fueled diesel engines.
ignited. Compared to the single RIF model, the multiple RIF model predicts the much higher level of the averaged conditional scalar dissipation rate. In context with this multiple RIF approach, the calculation starts with a single flamelet, and the first flamelet is split into two such that one of them represents the larger and the other represents the lower values of the scalar dissipation rate. Since this splitting process continues, it is quite possible that one of the flamelets can be ignited with a relatively low scalar dissipation rate. This is the main reason for the earlier ignition predicted by the multiple RIF approach. Figure 4 shows the instantaneous contours of mean temperature for 7° CA injection timing. Numerical results indicate that a large portion of the liquid droplets are still remaining at the late stage of injection, 14° CA, and these remaining droplets are nearly evaporated roughly at 16° CA. At 18° CA, the autoignition occurs and the partially premixed flames propagate the lean and rich mixture side in the downstream and upstream spray field. Validation Case II: DME Fueled Diesel Engines. In order to validate the predictive capability of the present turbulent combustion model for the spray combustion processes of the DME fueled diesel engines, the RIF model with the vaporization effects on turbulent combustion has been applied to simulate the spray dynamics, vaporization, autoignition,
combustion process, and pollutant (NOx and soot) formation in direct-injection DME fueled diesel engines. We have chosen the validation case as the measurements performed by Kim et al.27 The specification and computational conditions for the DME fueled diesel engine used in the experiment27 are given in Tables 3 and 4. The orifice diameter is 128 μm, and the spray inclination angle is 78°. Computations are performed for a 60° sector of the combustion chamber with periodic boundary condition in the azimuthal direction, and the spatial grid system at top dead center is presented in Figure 5. According to measurements,27 the starting time of energizing was used instead of the injection time and the injection delay time corresponding to the starting time of energizing was obtained through experiments. In the present numerical study, the injection time used as the input is estimated from the injection delay time evaluated from measurements. In Figure 6, the simulated cylinder pressure histories obtained by the single RIF model are compared with the measured data27 for the injection pressures, 35 MPa. The predicted ignition delay times are sligtly overetimated, compared to experimental data. However, the cylinder pressure are considerably overepredicted. In order to eliminate the potential errors arising from the neglect of the spatial inhomogeneity of scalar dissipation rate, the multiple RIF with 10 flamelets has been to analyze the spray combustion processes of the DI DME fueled diesel engine. As would be expected, the multiple RIF model predicts slightly earlier autoignition than the single RIF model. When the spatial
(27) Kim, M. Y.; Bang, S. H.; Lee, C. S. Experimental Investigation of Spray and Combustion Characteristics of Dimethyl Ether in a Common-Rail Diesel Engine. Energy Fuels 2007, 21 (2), 793–800.
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Figure 12. Instantaneous contours of NO concentration for 2° ATDC injection timing in DME and n-heptane fueled diesel engines.
inhomogeneity of scalar dissipation rate are included through the multiple RIF procedure, the predicted ignition delay time agrees reasonably well with experimental data. However, the cylinder pressure after autoignition is still noticeably overestimated. In order to identify the cause of these discrepancies, the instantaneous contours of mean temperature and the DME spray trajectories are displayed in Figure 7. In this relatively small DME fueled diesel engine, it can be clearly seen that the turbulent lifted flame over the highly evaporating DME spray jets is easily impinged on the cylinder wall. This flame/wall impingement with the high convective cooling and the high scalar dissipation rate results in the local extinction which directly leads to the decrease of cylinder pressure. However, the present RIF turbulent combustion model is unable to realistically simulate the local extinction process encountered in the flame/wall impingement situation. Therefore, these overestimated cylinder pressures are mainly attributed to the defects of the present RIF turbulent combustion model which is unable to realistically predict the local extinction process encountered in the flame/wall impingement situation. It is also speculated that this overestimated cylinder pressure could be partially caused by the neglect of the residual gases in the real engine operation. Comparison of Combustion Characteristics for n-Heptane and DME Fueled Diesel Engines. In the present study, in order to identify the distinctive differences of the spray
combustion characteristics in the DME fueled and n-heptane fueled DI diesel engine, computations utilizing the RIF turbulent combustion model to account for the vaporization effects are made for Cummins N-14 engine with the same engine specification and computational conditions given in Tables 1 and 2. Since the densities of the DME and n-heptane are slightly different, for the purpose of the parametric numerical study, the same amount of mass at the same injection duration is injected for two liquid fuels. In Figure 8, the simulated cylinder pressure and temperature histories for the DME and n-heptane fueled diesel engines are compared for three different injection timings. In terms of the cylinder pressure, compared to the n-heptane fuel, the DME fuel yields a much lower rise of pressure and volume-averaged temperature after the ignition for three injection timings because of its much lower heating value. Figure 9 shows the temporal evolutions of the scalar dissipation rate and the maximum temperatures of the flamelet for three injection timings of two liquid fuels. Numerical results clearly indicate that the ignition delay times increase with the injection time for both fuels, mainly owing to the much slower evaporation in the lower ambient temperature of the retarded injection period. Three injection timings of 2.0°, 7.0°, and 9.5° CA lead to the ignition delay times, 3.89°, 9.87°, and 13.72° CA for the DME fuel injection, and 6.40°, 16.18°, and 21.28° CA for the n-heptane fuel injection, respectively. Compared to the n-heptane liquid 4924
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Figure 13. Instantaneous contours of soot volume fraction for 2° ATDC injection timing in DME and n-heptane fueled diesel engines.
fuel, the DME liquid fuel yields the much shorter ignition delay times for three injection timings. By retarding the injection timing, this trend is progressively apparent. This implies that the n-heptane fuel yields much longer ignition delay times in the low-temperature and low-pressure environment corresponding to the retarded injection cases, compared to the DME fuel. These distinctly shorter ignition delay times of the DME fuel are mainly due to the much faster evaporation characteristics as well as the much shorter ignition delay times of the homogeneous mixture. At the initial stage of liquid fuel injection for the DME and n-heptane liquid fuels, the much higher interphase relative velocity results in the higher vaporization rate, the increase of the primary and secondary liquid atomization, and the enhanced turbulence intensity. Consequently, the higher interphase relative velocity at the early injection period abruptly increases the scalar dissipation rate by elevating these nonequilibrium effects such as the enhanced turbulence, the high spray evaporation rate, and the large gradient of mixture fraction. Compared to the n-heptane fuel, the DME fuel yields the much higher peak scalar dissipation rate due to its faster evaporation and short spray penetration characteristics. Shortly after this initial injection period, the scalar dissipation rate gradually decreases to the slow-varying low value since the fuel vapor spreads out the ambient flow field and the turbulence is continuously dissipated. At the small scalar dissipation rates, the diffusive losses become
so small that heat and radicals produced by the chemical reaction are continuously increased and the mixture field can be finally ignited. These numerical results also indicate that, after ignition, the DME fuel generates the noticeably higher mean scalar dissipation rate. This implies that the DME spray combustion at the high-temperature oxidation period could be more sensitive to the nonequilibrium effects on the turbulence-chemistry interaction, compared to the n-heptane spray combustion. Figures 10-13 show the instantaneous distribution of temperature, droplet trajectories, OH mass fraction, NO concentration, and soot volume fraction for the early injection timings (2° CA) in the n-heptane and DME fueled DI diesel engines. It can be clearly seen that the DME sprays have a relatively short spray penetration length due to their much higher evaporation rate, compared to the n-heptane sprays. For 2° CA injection timing, after the autoignition of the DME sprays occur at 3.89° CA, the spray penetration length is drastically decreased and the DME spray is nearly invisible after 6.0° CA. These numerical results suggest that, after this injection stage, 6.0° CA, the DME spray is quickly atomized and evaporated as soon as the DME liquid fuel is injected into the combustion chamber. On the other hand, the n-heptane spray before ignition maintains a fairly long spray penetration length. After autoignition at 6.40° CA, the penetration length of the n-heptane sprays is substantially decreased, but it maintains a certain 4925
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Figure 14. Time evolutions of temperature and mass fractions of various species in the interactive flamelet for n-heptane (left side)/DME (right side) spray (2° ATDC injection timing).
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Figure 15. Time evolutions of temperature and mass fractions of various species in the interactive flamelet for n-heptane (left side)/DME (right side) spray (7° ATDC injection timing).
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Figure 16. Time evolutions of temperature and mass fractions of various species in the interactive flamelet for n-heptane (left side)/DME (right side) spray (9.5° ATDC injection timing).
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: DOI:10.1021/ef9004016
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level during the main combustion processes. Since the DME sprays have a relatively short spray penetration length and a higher evaporation rate than the n-heptane sprays, the DME fuel creates a relatively narrow high-temperature flame zone relatively close to the injector. This unique trend in the DME fueled diesel engines is encountered for all injection timings. In the case of the DME sprays, the OH radical is stated to be formed at the much earlier injection period around at the autoignition stage, and it is highly distributed in the proximity of the relatively narrow hot flame zone. In case of the nheptane sprays, the OH radical is initiated to be created at the relatively late injection period around at the autoignition stage, and the relatively high OH concentration zone is formed at the relatively broad hot flame zone. Compared to the n-heptane sprays, the DME sprays yield a much higher peak OH level. This could be related to the characteristics of the oxygenated fuel. In DME fueled diesel engines, the peak OH levels for three injection timings occur near the fuel injector because of the relatively poor mixing process corresponding to the much shorter penetration length of the DME sprays. On the other hand, in the n-heptane fueled diesel engines, the OH radical is broadly distributed in the combustion chamber. In the case of the injection timing (2° CA) corresponding to the relatively high-temperature and highpressure environment, the peak OH level occurs near the fuel injector, similar to the DME fueled engines. However, in case of the late injections (7° and 9 °CA) corresponding to the relatively low-temperature and low-pressure environment, the peak OH level occurs at regions far from the fuel injector. These numerical results also indicate that the NO is distributed around at the hot flame zone and the n-heptane sprays generate a much broader NO distribution zone with a much higher concentration level. In the DME and n-heptane fueled diesel engines for three injections, it can be clearly seen that the NO is similarly distributed to the OH radical. This is mainly because the peak OH level occurs at the peak temperature region where the thermal NOx is mostly formed. For the DME and n-heptane sprays, the soot volume fraction is distributed at the fuel-rich regions. Compared to the nheptane fuel, the DME fuel generates a remarkably low soot level. Figures 14-16 present the time evolution of local flame structure in terms of temperature, mass fractions of the major and minor species, and the consumption rates of the oxygen near autoignition of n-heptane and DME fueled diesel engines for all injection timings (2°, 7°, and 9.5° CA), respectively. Similar to the temporal evolution of local flame structure in the constant volume chamber, after a sufficient time delay, the rapid decay of scalar dissipation rate allows the accumulation of heat and radicals, which can be found somewhat widely near the slightly rich mixtures close to the autoignition delay times. Thereafter, the temperature rapidly increases at a relatively narrower region in fuel lean mixtures, and the oxygen and fuel near stoichiometry are rapidly consumed; the temperature and mass fraction of the combustion products such as CO2 and H2O are highly increased. The mass fractions of H2 and CO are progressively increased according to the flame propagation in the fuel-rich region, and their peak levels occur at the fuel-rich region where the fuel oxidation rate is the highest. As a result of subsequent flame propagation toward both fuel-rich and fuel-lean regions, the reaction zone in the mixture fraction space becomes broader and the profiles of temperature and participating species approach a certain limit. After a rela-
tively long period, reactants within the fuel-lean side are almost burned out while the flame field on the fuel-rich side is still propagating. However, compared to the constant volume chamber situation, the ambient temperature and pressure at the engine situations are dependent on the injection time, the ignition delay time, and the piston movement. At nearly the end of the combustion period (130° CA) where the ambient temperature and pressure are considerably lower, the peak level of temperature and H2 and CO mass fractions are noticeably decreased. Numerical results also indicate that the increased injection timings in the DME fueled diesel engines result in a slightly broad flame zone in the fuel-rich region. This trend could be tied with the relatively high scalar dissipation rate obtained for the late injection case. In the overall flame structure, compared to the n-heptane fueled diesel engine, the DME fueled diesel engine has a much broader hot-temperature flame zone in the fuel-rich side of the mixture fraction space. This distinctly different structure of the DME spray flame could be mainly related to the characteristics of the oxygenated fuels. Again, the predicted profiles of oxygen mass fraction, fuel mass fraction, and oxygen consumption rate for both fuels clearly reflect this trend. In terms of the OH mass fraction, DME yields the much broader distribution at the fuel-rich region. Moreover, in the fuel-rich region, DME generates the much broader and higher hydrogen distribution which greatly reduces the soot formation in the actual spray flames. These numerical results suggest that the distinctly broader and higher OH and H2 distributions in the fuel-rich region can remarkably reduce the soot formation in the DME fueled diesel engines, compared to the conventional hydrocarbon fueled diesel engines. 4. Conclusions In the present study, in order to understand the overall spray combustion characteristics of DME fuel as well as to identify the distinctive differences of DME combustion processes against the conventional hydrocarbon liquid fuels, the sequence of the comparative analysis has been systematically made for DME and n-heptane fueled DI diesel engines. On the basis of numerical results, the following conclusions are drawn in terms of the spray combustion processes, flame structure, turbulence-chemistry interaction, and pollutant formation in the n-heptane and DME fueled engines. (1) In the numerical simulation of combustion processes in the DI n-heptane diesel engine, the computed and measured pressure histories and ignition delay times have reasonably good agreement for all cases. However, the overestimations in the ignition delay time and cylinder pressure are quite noticeable for the relatively late injection timings (7° and 9.5° CA). These discrepancies could be mainly related to the use of the nheptane fuel to simulate the multicomponent diesel fuel which has the different characteristics in vaporization, chemical kinetics, spray penetration, and two-phase air-fuel mixing processes. (2) When the spatial inhomogeneities of the scalar dissipation rate are included through the multiple RIF procedure, the predicted ignition delay time and cylinder pressure history have relatively good conformity with experimental data, compared to the single RIF approach. (3) In the numerical simulation of combustion processes in the DI DME fueled diesel engine, the predicted ignition delay times agree reasonably well with experimental data for the injection pressure, 35 MPa. However, the cylinder pressure 4929
Energy Fuels 2009, 23, 4917–4930
: DOI:10.1021/ef9004016
Yu and Kim
are considerably overestimated. When the spatial inhomogeneity of scalar dissipation rate are included through the multiple RIF procedure, the predicted ignition delay time is well agreed with experimental data, compared to the single RIF approach. However, the cylinder pressure after autoignition is still noticeably overestimated. These overestimated cylinder pressure are mainly attributed to the defects of the present RIF turbulent combustion model which is unable to realistically predict the local extinction process encountered in the flame/wall impingement situation. It is also speculated that this overestimated cylinder pressure could be partially caused by the neglect of the residual gases in the real engine operation. (4) With the same amount of the injected mass at the same injection duration, in terms of the cylinder pressure, compared to the n-heptane fuel, the DME fuel yields the much lower pressure rise after the ignition for three injection timings because of its much lower heating value. Compared to the n-heptane liquid fuel, the DME liquid fuel yields much shorter ignition delay times for three injection timings. These distinctly shorter ignition delay times of the DME fuel are mainly due to the much faster evaporation characteristics as well as the much shorter ignition delay times of homogeneous mixture. Compared to the n-heptane fuel, the DME fuel yields a much higher peak scalar dissipation rate due to its faster evaporation and short spray penetration characteristics. These numerical results also indicate that, after ignition, the DME fuel generates a noticeably higher mean scalar dissipation rate. This implies that the DME spray combustion in the high-temperature oxidation period could be more sensitive to the nonequilibrium effects on the turbulence-chemistry interaction, compared to the n-heptane spray combustion. (5) The increased injection timings in the DME fueled diesel engines result in the slightly broad flame zone in the fuel-rich region, and this trend could be tied with the relatively high scalar dissipation rate obtained for the late injection case. In terms of the overall flame structure, compared to the n-heptane fueled diesel engine, the DME fueled diesel engine has the much broader hot-temperature flame zone in the fuel rich side of the mixture fraction space mainly due to the
characteristics of the oxygenated fuels. In terms of the OH mass fraction, DME yields a much broader distribution at the fuel-rich region. Moreover, in the fuel-rich region, DME generates a much broader and higher hydrogen distribution which greatly reduces the soot formation in the actual spray flames. (6) It can be clearly seen that the DME sprays have a relatively short spray penetration length due to their much higher evaporation rate, compared to the n-heptane sprays. After the autoignition of the DME sprays occurs at 3.89° CA, the spray penetration length is drastically decreased and the DME spray is nearly invisible after 6.0° CA. These numerical results suggests that, after this injection stage, 6.0° CA, the DME spray is quickly atomized and evaporated as soon as the DME liquid fuel is injected into the combustion chamber. Since the DME sprays have a relatively shorter spray penetration length and a higher evaporation rate than the n-heptane sprays, the DME fuel creates a relatively narrow high-temperature flame zone relatively close to the injector. (7) In the case of the n-heptane fuel, the soot is rapidly formed at the fuel-rich and relatively hot zone during the initial stage of combustion, and the soot level at the late stage of combustion is gradually decreased by the soot oxidation process which is mainly contributed by oxygen and OH radicals transported by the turbulent mixing process. By retarding the fuel injection, the n-heptane sprays yield a slightly lower soot emission level. This trend might be related to the relatively longer spray penetration and the lower soot formation rate in the relatively low ambient pressure and temperature environment corresponding to the retarded injection. In case of the DME fuel, the relatively low level of soot is formed during the initial stage of combustion, and the produced soot at the late stage of combustion is nearly oxidated. Thus, the soot emission levels for three injection timings are virtually zero in this DME fueled diesel engine. Acknowledgment. This work was supported by the CEFV (Center for Environmentally Friendly Vehicle) of the Eco-STAR project from MOE (Ministry of Environment, Republic of Korea).
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