Energy Fuels 2010, 24, 2429–2438 Published on Web 03/08/2010
: DOI:10.1021/ef901211d
Numerical Investigation of Combustion and Exhaust Emissions Characteristics Based on Experimental Spray and Atomization Characteristics in a Compression Ignition Diesel Engine Su Han Park,† Hyung Jun Kim,‡ and Chang Sik Lee*,‡ †
Department of Mechanical Engineering, Graduate School of Hanyang University, 17 Haengdang-dong, Sungdong-gu, Seoul, 133-791, Korea, and ‡Department of Mechanical Engineering, Hanyang University 17 Haengdang-dong, Sungdong-gu, Seoul 133-791, Korea Received October 24, 2009. Revised Manuscript Received February 22, 2010
The aim of this study is to investigate the effect of fuel injection pressure and timing on the combustion and exhaust emissions characteristics in a compression ignition diesel engine. For these investigations, a study on the spray and atomization characteristics was performed. The spray and atomization characteristics were experimentally analyzed in terms of spray tip penetration, spray cone angle, droplet size, and velocity. On the basis of the spray-atomization characteristics, the combustion and exhaust emissions characteristics were calculated using a KIVA-3 V code. The numerical analysis (KIVA-3 V) was applied for the acquisition of the calculated combustion and exhaust emission characteristics after the verification of calculated results by the experimental results obtained from a single cylinder diesel engine. Measurements revealed that the spray development with high injection pressure is faster, and the spray tip penetration becomes longer compared to those with low injection pressure. At the initial spray stage, the high pressure injection spray has a high overall axial velocity, resulting in good atomization of droplets with high velocity. In addition, the spray droplet size decreased as the injected spray decreased. On the other hand, when the injection timing retarded to around top dead center (TDC), the peak combustion pressure and the peak rate of heat release decreased. When the injection pressure increased, the soot, HC, and CO emissions decreased, but NOx emissions increased somewhat. In addition, the retarded injection timing caused a decrease in NOx emissions, while soot, CO, and HC emissions increased.
during the combustion process and homogenization of the in-cylinder air-fuel mixture. Ochoterena et al.5 studied the spray behavior and combustion characteristics of alternative fuels as well as diesel fuel. They discovered that high viscosity fuels have wider spray cone angles, smaller discharge coefficients, and shorter vapor penetration than low-viscosity fuels. In addition, they revealed that the particulate matter emissions are generally lower when alternative fuels are used. Suh et al.6 reported on the spray and combustion characteristics as a function of the blending ratio of diesel and biodiesel fuels under various injection conditions, such as single and pilot injections. They showed that the atomization and combustion performances were improved by using pilot injection, and CO and HC emissions decreased as a result of the enhanced fuel atomization. Further, they found that NOx emissions increased for higher combustion chamber temperatures, caused by heat released when the pilot injection timing approaches the main injection timing. Like these,4-6 many investigations on diesel spray and combustion characteristics are actively progressing. However, a major portion of investigations still deals with an independent study of spray and combustion characteristics, respectively. Therefore, in this work, the analytical study on the effect of the diesel spray on the combustion and exhaust emissions characteristics were performed based on the experimental study of diesel spray characteristics.
1. Introduction Despite the many advantages of diesel engine such as the high thermal efficiency and high power performance, there still remain some problems relating to exhaust emissions of nitrogen oxides (NOx), hydrocarbons (HCs), carbon monoxide (CO), and particulate matter (PM). In particular, it is known that compression ignition diesel engines produce higher NOx and soot emissions compared to spark ignition engines. Many researchers have been working to find alternative combustion mode1 and injection strategies.2,3 Fuel spray characteristics play an important role in the combustion characteristics and exhaust emissions formation process in a direct injection diesel engine in which the fuel spray is directly injected into the combustion chamber. In addition, the quality of the fuel-air mixture formation is also a very important factor in reducing emissions. Therefore, the spray and atomization characteristics have to be considered to reduce exhaust emission and improve the combustion performance. Yun et al.4 revealed that low NOx and soot emissions are achieved through minimization of peak temperatures *To whom correspondence should be addressed. Telephone: þ82-22220-0427. Fax: þ82-2-2281-5286. E-mail:
[email protected]. (1) Park, S. W.; Reitz, R. D. Combust. Sci. Technol. 2007, 179, 2279–2307. (2) Reitz, R. D. Combust. Sci. Technol. 1998, 138, 257–278. (3) Pierpont, D. A.; Montgomery, D. T.; Reitz, R. D. SAE Technical Paper SAE 950217, 1995. (4) Yun, H.; Sellnau, M.; Milovanovic, N.; Zuelch, S. SAE Technical Paper SAE 2008-01-0639, 2008. r 2010 American Chemical Society
(5) Ochoterena, R.; Larsson, M.; Andersson, S.; Denbratt, I. SAE Technical Paper SAE 2008-01-1393, 2008. (6) Suh, H. K.; Roh, H. G.; Lee, C. S. J. Eng. Gas Turbines Power 2008, 130, 032807.
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Figure 1. Schematic of the injection rate measuring system.
Also, in the numerical analysis of diesel combustion, the combustion and exhaust emissions results in a coarse mesh were compared to those in a fine mesh in order to investigate the mesh size dependence of diesel combustion and exhaust emissions performance. The spray and atomization characteristics were analyzed in terms of the spray tip penetration, spray cone angle, and Sauter mean diameter (SMD) in a constant volume chamber. On the basis of the spray and atomization characteristics, the calculated combustion and emission characteristics such as the combustion pressure, rate of heat release, and exhaust emissions were analyzed. 2. Experimental Setup and Procedure 2.1. Experimental Setup. Figure 1 shows the schematic diagram of the injection rate measuring system. The injection rate measuring system consists of the fuel injection part and the data acquisition part. This apparatus is based on the measurement of the pressure variation of a tube filled with fuel when the fuel is injected to the tube.7 During the experiment, the pressure in the tube was set equal to 3.0 MPa. A piezo type pressure sensor was utilized to measure the pressure of the tube, and the injection quantity was measured by the average of 5 000 continuous injections for each test case. Injection rate profiles obtained from the injection rate measuring system were applied as input data for the numerical method used to calculate the diesel combustion and exhaust emissions characteristics. In order to visualize and analyze the diesel spray, a spray visualization system and droplet measuring system were installed as shown in Figure 2. In both systems, the diesel fuel was supplied through a common rail injection system with a high pressure pump and six hole solenoid type injector with a 0.128 mm hole diameter. The injected high pressure diesel spray was visualized using a high speed camera (FASTCAM APX-RS, Photron) with a metal-halide lamp. The spray images were digitally captured using an image grabber. In addition, the injection signal from the injector driver (TDA-3200H, TEMS) and the shutter signal of the high-speed camera were synchronized by a digital delay/pulse generator (model 555, Berkeley Nucleounics Corp.). The spray development and behavior of diesel fuel are observed in a constant volume chamber. The spray behavior in the combustion cylinder is mainly affected by the ambient gas density, not an independent parameter such as temperature or pressure.8 Because the ambient gas density in the engine cylinder is close to 20 kg/m3 around top dead center (TDC), the ambient pressure in the chamber is fixed as 2 MPa in order to reproduce engine cylinder conditions. A droplet measuring system (phase Doppler particle analyzer, PDPA) shown in Figure 2a was used to analyze the droplet
Figure 2. Schematic of the spray visualization system and the droplet measuring system (a) and the measuring points for the analysis of the diesel droplets (b). Table 1. Criteria for Droplet Measurement System (PDPA) burst threshold (mV) mixer frequency (MHz) filter frequency (MHz) PMT voltage (V) signal-to-noise ratio diameter subrange (μm)
0.5 36 40 500 65 2-75
Table 2. Test Engine Specifications test injector
single cylinder diesel engine
number of nozzle holes nozzle diameter spray angle engine type bore stroke displacement volume compression ratio intake valve close timing exhaust vale open timing
6 0.128 mm 156° direct injection diesel engine with a natural aspiration type 75.5 mm 84.5 mm 373.3 cm3 17.8:1 BTDC 128° ATDC 172°
size and velocity of the diesel fuel. This system consists of an Ar-ion laser (INNOVA 70C, Coherent), a transmitter, a receiver, and a signal analyzer. In this work, the microscopic spray characteristics such as droplet size and velocity are measured under the atmospheric conditions, contrary to experiments on the macroscopic spray characteristics. This is due to the uncertainty of the obtained data. When the laser passes through the windows, it is reflected. Although many researches are attempting to improve the accuracy of data in high ambient pressure conditions, there still remains some uncertainty. Also, in this study, the optimal conditions based on the data acquisition rate and signal intensity are
(7) Bosch, W. SAE Technical Paper SAE 660749, 1966. (8) Pickett, L., M.; Kook, S.; William, T. C. SAE Technical Paper SAE 2009-01-0839, 2009.
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Figure 3. Schematic of the test engine for the validation of the diesel combustion and exhaust emission characteristics (a) and the calculation coarse grid (b) and fine grid (c).
flow in the crevice region. Figure 3b and 3c show the calculation grids.
applied to the PDPA setup for the improvement of the experimental accuracy and reliability.9,10 Therefore, the laser output and PMT (photomultiplier tube) voltage were set to 700 mW and 500 V, respectively. The effective range of droplet diameter in the PDPA signal analyzer was from 2 to 75 μm considering that the nozzle diameter was 0.128 mm, approximately 10 000 droplets were collected and averaged at each measurement point illustrated in Figure 2b. When fuel droplets pass through the measurement volume, the dispersed Ar-ion laser beam is detected at the receiver. The detected phase difference and Doppler signal frequency are then converted to the mean droplet size and velocity. Detailed specifications of the PDPA system are shown in Table 1. 2.2. Test Engine Specification and Computational Grid. In order to validate the numerical results, the experimental results for the standard case, including the combustion pressure and the rate of heat release, were obtained from a single cylinder diesel engine which is further detailed in Table 2 and Figure 3a. The compression ratio of the test engine was 17.8 to 1 and the combustion chamber was a reentrant-type. The test engine valve train was a double overhead cam type, and the valve mechanism had two intake valves and two exhaust valves. In the numerical study, a 60° sector mesh of the combustion chamber considering the fuel injector with six holes was used to calculate the combustion and exhaust emissions characteristics. Computation cells in the cylinder bowl were created using three parts, and the crevice volume was used for the fuel
3. Numerical Model for the Combustion and Exhaust Emission Characteristics Analysis In order to analyze the combustion and emission characteristics as functions of injection pressure and timing, calculations were carried out using the KIVA-3 V code. In this study, a tetradecane (C14H30) was utilized for the analysis of spray and combustion characteristics because this fuel has properties similar to those of conventional diesel fuel. The Kelvin-Helmholtz and Rayleigh-Taylor hybrid breakup model11 was applied to calculate the spray and atomization characteristics of injected droplets. Also, turbulent flow in the cylinder and the spray development process after impingement on the cylinder wall were calculated using the renormalization group (RNG) k-ε model12 and the spray-wall interaction model.13 In the ignition and combustion of diesel fuel, a reduced Engine Research Center (ERC) n-heptane mechanism14 composed of 36 species and 74 reactions was applied in the KIVA-3 V code. Regarding the n-heptane reaction mechanism, Lawrence Livermore National Laboratory (LLNL) developed it for detailed chemical kinetic (11) Beale, J. C.; Reitz, R. D. Atomization Sprays 1999, 9, 623–650. (12) Han, Z.; Reitz, R. D. Combust. Sci. Technol. 1995, 106, 267–295. (13) O’Rourke, P. J.; Amsden, A. A. SAE Technical Paper SAE 200001-0271, 2000. (14) Patel, A.; Kong, S. C; Reitz, R. D. SAE Technical Paper SAE 2004-01-0558, 2004.
(9) Araneo, L.; Soare, V.; Payri, R.; Shakal, J. J. Phys., Conf. Ser. 2006, 45, 85–93. (10) Lacoste, J.; Crua, C.; Heikal, M.; Kennaird, D.; Gold, M. SAE Technical Paper SAE 2003-01-3085, 2003.
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Figure 4. Overall chemical oxidation mechanisms for n-heptane.
model including 250 species and 576 reactions15 as shown in Figure 4. In order to reduce the computational time for the ignition and combustion process, a reduced reaction mechanism for n-heptane was used to simulate the two stage ignition of diesel oxidation. The Chemkin II chemistry solver16 was integrated into the KIVA code and was used to calculate the chemical reactions occurring during combustion. In order to predict the NOx and soot emissions, the reduced GRI NO mechanism (four species (N, NO, N2O, NO2) and nine reactions) and a two-step phenomenological soot model17 were used. The Hiroyasu soot formation18 and the Nagle-Strickland-Constable (NSC) oxidation model19 were used to apply the two-step phenomenological soot model in this study. The change rate of soot mass was calculated from the discrepancy between the rate of formation and the rate of oxidation as the following equation. dMsoot dMformation dMoxidation ¼ ð1Þ dt dt dt In this equation, the formation rate of soot mass was calculated from the equation of Arrhenius form multiplied by fuel vapor mass (Mfv) as shown in the eq 2. In addition, the soot oxidation rate was calculated by using the NSC oxidation model based on an empirical equation from the oxidation experimental results of the carbon graphite as given by the following equations.20 dMformation Ef 0:5 ¼ Af Mfv Plocal ð2Þ exp dt RT dMoxidation 6MWcarbon ¼ Msoot Rtotal dt Fsoot Dsoot
Figure 5. Comparison between the experimental and calculated combustion pressure and rate of heat release characteristics (symbol — experimental result; line — calculation result, Pinj = 50 MPa, 100 MPa, mfuel = 10 mg, injection timing = BTDC 8°).
Figure 6. Injection rate profile of diesel fuel.
ð3Þ
For analyzing the other emissions (e.g., CO and HC), additional emission models were not needed due to the fact that these species are included in the reduced ERC n-heptane mechanism. 4. Results and Discussion 4.1. Model Validation. The experimental combustion pressure and rate of heat release (ROHR) at an injection timing of BTDC 8° was selected for the validation of the calculation results. The naturally aspirated single cylinder diesel test (15) Curran, H. J.; Gaffuri, P.; Pitz, W. J.; Westbrook, C. K. Combust. Flame 1998, 114, 149–177. (16) Kee, R. J.; Rupley, F. M.; Miller, J. A. Sandia Report SAND 898009, 1989. (17) Kong, S. C.; Sun, Y.; Reitz, R. D. J. Eng. Gas Turbine Power 2007, 129, 252–260. (18) Hiroyasu, H.; Katota, T. SAE Technical Paper SAE 760129, 1976. (19) Nagle, J. S.; Constable, R. F. Proceedings of the Fifth Carbon Conference 1962, 1, 154. (20) Han, Z.; Uludogan, A.; Hampson, G. J.; Reitz, R. D. SAE Technical Paper SAE 960633, 1996.
Figure 7. Diesel spray development process according to the elapsed time after the start of energizing (Pinj = 50 MPa, 100 MPa, Pamb = 2 MPa, mfuel = 10 mg).
engine was operated at the following conditions: injection pressures of 50 MPa and 100 MPa, an injected fuel quantity per injection of 10 mg, and an injection timing of 8° before top dead center (BTDC). Figure 5 shows the comparison 2432
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Figure 8. Spray tip penetration and spray cone angle characteristic characteristics (Pinj = 50 MPa, 100 MPa; Pamb = 2 MPa; mfuel = 10 mg; black empty symbol — spray tip penetration; colored solid symbol — spray cone angle).
Figure 10. Comparison between two different injection pressures on the overall axial velocity and tip velocity characteristics (Pamb = 0.1 MPa, mfuel = 10 mg).
calculated and experimental ROHR, the calculated ROHR is quite higher than the experimental ROHR. This results from a difference in calculation procedure. The heat release rates of experimental results are acquired from the simple thermodynamic analysis which assumes the conservation of energy with the measured in-cylinder pressures. Even though heat loss to the cylinder wall and mass loss (blow-by) to the crevice volume, etc. are generated in the combustion chamber, a simple thermodynamic equation does not consider these terms. On the other hand, numerical models predict the higher heat release rate compared with the measured data, and it results from the different detecting procedures. In the numerical models, heat release rate is directly calculated from the chemical reaction of fuels, and thus it has a higher value because it does not consider the heat losses after reactions. Also, the wall heat loss model for the calculation used was a base model without modification of the subroutine in the KIVA-3 V code. 4.2. Injection and Overall Spray Characteristics. Figure 6 shows the diesel spray injection rate profile obtained from
Figure 9. Diesel spray atomization characteristics.
between the experimental and calculated combustion pressure and ROHR. The calculated results show a good agreement with the calculated results. In the comparison of the 2433
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Figure 11. Combustion pressure and rate of heat release characteristics of diesel fuel according to the injection timing at different injection pressures.
the injection rate measuring system. The peak injection rate, the injection delay, and the real injection duration were analyzed from the injection rate profile. The injection delay is defined as the time interval between the start of energizing and the initial increase in the injection rate; the real injection duration is defined as the period over which fuel is delivered through the injector indicated by the injection rate, not the energizing duration. At a fixed injection quantity (mfuel = 10 mg), the increased injection pressure induced an increase in the peak injection rate, resulting in a decrease in the real injection duration. This indicates that high pressure in the mini-sac of the injector causes high injection velocity at the early stage of injection duration. In addition, the injection delay becomes short with an increase of the injection pressure. For high injection pressure, the injection rate profile suddenly increased and decreased, while it gradually increased and decreased for low injection pressure. These injection characteristics follow previous research.8 This is because an increase in the injection pressure causes the rapid opening of the needle and influences the injection delay and real injection duration.21 It is also expected that this injection characteristic affects the spray development process and atomization characteristics.
Figure 7 shows the diesel spray development process as a function of elapsed time after the start of energizing. As mentioned in the previous paragraph, the diesel spray development at high injection pressure is faster than that at low injection pressure at the same time because the injection rate profile suddenly increased at high injection pressure. In addition, it seems that the spray tip penetration for high injection pressure is longer than that in a low injection pressure. At 0.6 ms after the start of energizing, the spray image of the high injection pressure is clearer than that of the low injection pressure. This result can be explained from the analysis of the injection rate profile, i.e., the injected quantity at high injection pressure is larger than that at low injection pressure at the same time (tasoe = 0.6 ms). The spray tip penetration and spray cone angle based on the analysis of Figure 7 are illustrated in Figure 8. The spray tip penetration is defined as the maximum distance of the injected spray from the nozzle hole. The spray cone angle is the angle between two straight lines of the nozzle tip and maximum outer spray points. In this work, the spray tip penetration and cone angle is averaged from six plumes of injection nozzle. After the end of injection (Pinj = 50 MPa, around 1.1 ms; Pinj = 100 MPa, around 0.8 ms), the increased rate of spray tip penetration became blunt and the spray cone angle began to increase. This is the reason why
(21) Park, S. W. Ph.D. Dissertation, Hanyang University, Seoul, Korea, 2005.
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Figure 13. NOx and soot emission characteristics for the injection timing and injection pressure.
Figure 12. Temperature distributions in the combustion cylinder at two energizing timings and two injection pressures.
increasing intervals in both curves. In the case of high pressure injection, it can be observed that the increasing interval of SMD occurs earlier than the case of low-pressure injection. This is because there are active collisions and coalescence among droplets near the nozzle tip because droplet breakup time is short and atomization is improved by high-pressure injection. On the whole, the droplet size gradually decreased as the injected droplets progressed downward, as shown in Figure 9a. This result is explained by the droplet distribution as illustrated in Figure 9b, which shows the droplet number distribution as a function of the droplet size at 50 MPa of injection pressure. When the distance between the nozzle tip and the injected spray become large, the number of small droplets increased and the number of larger droplets decreased. In this work, the small droplet means the droplet with the size less than 15 μm, and the large indicates the droplet with the size more than 15 μm. Atomization performance is mainly affected by the relative velocity between the injected droplets and ambient air. Therefore, it is possible that the higher relative velocity induced better atomization performance.24 Figure 10 shows
the spray momentum rapidly decreased, and it was further affected by the aerodynamic force of the ambient gas.22 In comparison of the spray cone angle, the spray cone angle for high-pressure injection immediately after injection is lower than that for low-pressure injection due to the fast injection and large momentum. However, after the end of injection, the former is higher than the latter. It is believed that small droplets by the atomization and entrained ambient air easily spread to the outer region at high injection pressure. It is well-known that a higher injection pressure produces more effective air entrainment, in terms of entrainment quantity and rate.23 Figure 9 shows the droplets atomization characteristics such as local SMD and number of droplets. The local SMD is defined as the mean droplet size at the specific measuring points, while the number of droplets indicates the number of detected droplets with the same droplet size. With the comparison of the total mean droplet size, the increase in injection pressure from 50 to 100 MPa induced a reduction in droplet size by about 10%. In the droplet size distribution shown in Figure 9a, it is observed that there are two (22) Roisman, I. V.; Araneo, L.; Tropea, C. Int. J. Multiphase Flow 2007, 33, 904–920. (23) Choi, W.; Choi, B.-C. J. Automobile Eng. 2005, 129, 1025–1036.
(24) Lefebvre, A. H. Atomization Sprays; Hemisphere Pub. Corp.: New York, 1989; pp 274-275.
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velocity, while the tip velocity at 100 MPa shows a similar or little higher value than that at 50 MPa. This phenomenon can be explained as follows. The high pressure injection induced a good atomization as well-known. So, the spray droplet at high pressure injection is smaller than that at low pressure injection. Consequently, the overall axial velocity at 100 MPa of injection pressure with small droplets is low due to the drag force. On the other hand, the tip velocity mainly caused a fast tip velocity. From these results, it can be said that the high pressure injection spray has a fast tip velocity; however, the overall spray velocity is smaller than the low pressure injection spray due to the high drag force by a good atomization performance. 4.3. Calculation of Combustion and Exhaust Emissions Characteristics. The comparison between combustion characteristics such as combustion pressure and rate of heat release in the coarse and fine mesh according to the injection timing and injection pressure were illustrated in Figure 11. First of all, in the comparison according to the mesh type, the combustion characteristics in the coarse mesh show a similar pattern with those in the fine grid mesh. However, there are differences in the graph form of combustion pressure and heat release rate between the coarse and fine meshes. In the combustion pressure, the fine mesh shows a lower peak of combustion pressure after the ignition and longer combustion duration than the coarse mech. This is because the injected and evaporated fuel in the fine grid mesh has a wide spatial distribution by the effect of the active momentum transfer at the nearest cells. This led to the spread of the combustion region and long combustion duration. In addition, the heat release rate in the fine mesh shows the lower and longer patterns than that in the coarse mesh. On the other hand, the retardation of the injection timing causes a decrease in the peak combustion pressure as shown in parts a and b of Figure 11. In the analysis of ROHR, the ROHR at 50 MPa of injection pressure showed little decreasing trend according to the retardation of the injection timing. However, at 100 MPa of injection pressure, there is almost no difference of ROHR according to the injection timing except the injection timing of BTDC 16°. When the fuel injection timing is advanced, the injected diesel spray mixed well with ambient air in the combustion chamber, forming a good mixture during sufficient mixing time. With a good mixture in the combustion chamber, it is possible to acquire a high peak combustion pressure and ROHR through instant and uniform combustion. In the case of injection timing at BTDC 4° and BTDC 8°, combustion progressed in the expansion stroke by injection delay and ignition delay. When the injection pressure increased, the peak combustion pressure and ROHR also increased due to good atomization characteristics (Figure 9a). In addition, the ignition delay became short, and the increasing rate of the combustion pressure became steeper. For high combustion pressure, the ROHR also showed a high value. The in-cylinder temperature distribution and spray behavior in the coarse and fine meshes are illustrated in Figure 12 for two injection pressures and two injection timings. As shown in Figure 12a, in both calculation meshes, the high-temperature region in the combustion cylinder was widely distributed at the high injection pressure for the same injection timing. In addition, it was observed that the major part of the diesel combustion occurred at the piston bowl. In the case of the injection of BTDC 4°, the evaporated fuel flowed in the squish region. It can be expected that this
Figure 14. CO and HC emission characteristics for the injection timing and injection pressure.
the overall axial velocity and tip velocity of the injected diesel droplets at injection pressures of 50 and 100 MPa. The overall axial velocity is the average value which calculated from all of droplets at the same time after the energizing in each measuring points, and the tip velocity is the time derivative of the spray tip. As shown in Figure 10, just after injection, the overall axial velocity and tip velocity of the injected droplets at 100 MPa suddenly increased and showed the higher peak velocity compared to 50 MPa. In two injection pressures, the overall axial velocity and tip velocity showed the decreasing pattern according to the elapsed time after the start of energizing. In comparison between both injection pressures, it can be known that a high injection pressure observed a rapid decay of axial velocity compared to that of a low injection pressure. In addition, parts a and b of Figure 10 show a similar velocity variation trend, while the absolute value is some different. The velocity of Figure 10b was calculated at the spray tip, and then it shows the almost maximum velocity. In a detail comparison of both figures, the overall axial velocity at high pressure injection (Pinj = 100 MPa) shows somewhat lower value than that at the low pressure injection condition (Pinj = 50 MPa) after the peak 2436
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Figure 15. HC and NOx emission distributions in the combustion cylinder at two energizing timings.
characteristic affects a high HC emission because the combustion reaction is not active in the squish region due to low oxygen concentration. On the other hand, in the case of the injection of BTDC 16°, the almost injected spray after ignition still remained in the piston bowl. The temperature distribution in the coarse and fine meshes is different due to the difference in the spray progressing velocity. This characteristic can be explained by parts b and c of Figure 12. Parts b and c of Figure 12 show the calculated spray behavior for two injection timings in the coarse and fine meshes. As shown in both parts (Figure 12b,c), the spray development in the fine mesh is more advanced along the bowl surface than that in the coarse mesh because injected droplets have a fast velocity by the effect of dense cells with momentum regardless of the injection pressure and the injection timing. In addition, the spray in the fine mesh impacted against the right bowl wall before the spray in the coarse mesh at the same crank angle. From this reason, the wall wetting
in the fine mesh at the impingement point between the injected spray and piston bowl wall occurs and this leads to the deterioration of combustion and exhaust emission performance. NOx and soot emissions characteristics in the coarse and fine meshes are illustrated in Figure 13. The NOx and soot emissions in the fine mesh show a similar trend with those in the coarse mesh. However, NOx emission in the fine mesh shows a similar or lower amount compared to that in the coarse mesh, and soot emission in the fine mesh shows a higher value than that in the coarse mesh. In addition, the exhaust emissions of NOx and soot showed a trade-off relationship. The retardation of the injection timing induced a reduction of the NOx emission amount per power, while the soot emission amount increased. Increased injection pressure caused an increase in NOx emission and a decrease in the soot emission. These phenomena (correlation between NOx and soot) result from brisk combustion by the good atomization 2437
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performance. In addition, the temperature distribution with a high level in the combustion chamber (Figure 12) results in an increase in NOx emissions.25,26 Figure 14 shows CO and HC exhaust emissions at various injection timings. Both emissions increased with the retardation of the injection timing. As mentioned, Figures 11 and 12 show that the retardation of the injection timing caused low combustion pressure and combustion temperature. Low combustion temperature leads to a low oxidation rate of CO and suppressed consumption of hydrocarbon. As shown in Figure 14, the high injection pressure induced a reduction in emissions of CO and HC due to the active combustion resulting from good mixture formation and the wide high temperature distribution. On the other hand, the CO and HC emissions in the coarse and fine meshes have a similar tendency, but the HC emission in the fine mesh has higher values than that in the coarse mesh. This is why the spray in the fine mesh impinged on the right bowl wall before the spray in the coarse mesh at the same crank angle. The fuel wall-wetting on the bowl wall occurs and this advanced along the bottom bowl shape. Figure 15 shows HC and NOx emission distributions in the combustion cylinder. As shown in Figure 15, there are few HC emissions at the high injection pressure and the advanced injection timing due to the active combustion. Also, HC emissions appeared at the bowl bottom due to the wallwetting between the fuel and cylinder wall. However, at the same conditions (high injection pressure and advanced injection timing), NOx emissions and distributions were high and wide, respectively. This results from the high temperature and high combustion pressure in the combustion cylinder. Thus, the advanced injection resulted in a long duration of good mixture formation and then induced an instant explosion in the compression stroke.
the latter at the region after the end of injection. This is due to the aerodynamic drag of the ambient air. (2) The combustion results in a fine mesh and shows a lower peak combustion pressure and longer combustion duration than those in a coarse mesh. It is due to the injected and evaporated fuel in the fine grid mesh has wide spatial distribution by the effect of the active momentum transfer at the nearest cells. Also, the heat release rate in the fine mesh shows the lower and longer patterns than that in the coarse mesh. On the other hand, the spray development in the fine mesh is more advanced along the bowl surface than that in the coarse mesh because the injected droplets have a fast velocity by effect of dense cells with momentum. (3) In a fine mesh, NOx emission represented a similar or lower amount compared to that in the coarse mesh, and soot emission was a higher value than that in the coarse mesh. Also, the increased injection pressure caused an increase in NOx emission and a decrease in the soot emission due to the brisk combustion by the good atomization performance. (4) In two cases of fine and coarse meshes, the CO and HC emissions have a similar tendency, but the HC emission in the fine mesh has higher values than that in the coarse mesh. On the other hand, when the injection pressure increased, soot, HC, and CO emissions decreased but NOx emissions somewhat increased. In addition, the retarded injection timing caused a decrease in the NOx emission, while the soot, CO, and HC is increased. Therefore, it is expected that the reduction of exhaust emissions and the improvement of the atomization performance can be achieved through the high pressure injection and control of the injection timing. Acknowledgment. This work was supported in part by the CEFV (Center for Environmentally Friendly Vehicle) of the Eco-STAR project of the MOE (Ministry of the Environment in Seoul, Republic of Korea) and the Second Brain Korea 21 Project. This work also financially supported by a manpower development program for Energy & Resources supported by the Ministry of Knowledge and Economy (MKE).
5. Conclusions In this work, the spray and atomization characteristics of diesel fuel in a compression ignition diesel engine are experimentally investigated at various injection conditions. On the basis of these experimental spray characteristics, the combustion and exhaust emissions performance are studied using a numerical analysis method in two kinds of meshes with different grid sizes. From the results and discussions, the conclusions can be summarized as follows. (1) At constant injection quantity, the increased injection pressure induced an increase in a peak injection rate and a decrease of the real injection duration and injection delay. This is due to the shortening of the energizing duration by the increase of the injection pressure. From the analysis of the spray cone angle characteristics, the high pressure injected spray has a narrower spray cone angle than the low pressure injected spray at the initial stage after injection, while the former shows larger than
Nomenclature ATDC = after top dead center BTDC = before top dead center Dsoot = diameter of soot (cm) Ef = activation energy LZ = axial distance from the nozzle tip (mm) Msoot = soot mass Mformation = mass of soot formation Moxidation = mass of soot oxidation Mfv = fuel vapor mass MWcarbon = carbon molecular weight mfuel = injection quantity (mg) Pamb = ambient pressure (MPa) Pinj = injection pressure (MPa) Plocal = local pressure (MPa) R = universal gas constant Rtotal = net suface reaction rate SOE = start of energizing tasoe = time after the start of energizing (ms)
(25) Sher, E. Handbook of Air Pollution from Internal Combustion Engines Pollutant Formation and Control; Academic Press: New York, 1989. (26) Turns, S. R. An Introduction to Combustion: Concepts and Applications, international ed.; McGraw-Hill: Boston MA, 2000.
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