Energy & Fuels 2008, 22, 1581–1588
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Search for the Optimizing Control Method of Compound Charge Compression Ignition (CCCI) Combustion in an Engine Fueled with Dimethyl Ether Junjun Zhang,* Xinqi Qiao, Bin Guan, Zhen Wang, Guangei Xiao, and Zhen Huang Key Laboratory of Power Machinery and Engineering, Ministry of Education, Shanghai Jiao Tong UniVersity, Shanghai 200240, China ReceiVed December 22, 2007. ReVised Manuscript ReceiVed March 21, 2008
Taking advantage of HCCI combustion and in-cylinder direct-injection combustion for an engine fueled with dimethyl ether (DME), a new concept combustion system, namely, a compound charge compression ignition (CCCI) combustion system by port aspiration and direct injection of DME, is proposed. A diesel engine was modified to carry out this test. Combustion and emission characteristics were investigated in different premixed fuel ratios and fuel delivery advance angles. To further optimize the emissions and improve the DME fuel consumption in a CCCI engine, port fuel design concept and port aspiration of CO2 were employed. The experimental results show that the CCCI combustion process comprises HCCI combustion, premixing combustion, and diffusion combustion. The combustion characteristic is mainly decided by a premixed fuel ratio and CO2 concentration in air charge. In comparison to a homogeneous charge compression ignition (HCCI) combustion mode, CCCI combustion can extend the operating range with low NOx, hydrocarbon (HC), and CO emissions. To control the ignition and combustion phase of HCCI, the effect of liquid petroleum gas (LPG) added to DME fuel was evaluated. As a result, NOx emissions decreased and thermal efficiency was improved with the mix of LPG. It is of interesting to note that HCCI-MK (modulated kinetics) combustion for CCCI combustion engines was observed at a 14°CA before top dead center (BTDC) fuel delivery advance angle with an appropriate CO2 concentration in air charge, and NOx emissions could be lowered to near-zero levels.
1. Introduction Emission substances, particulate matter (PM) and NOx, generated from compression ignition engine create serious environmental problems. CO2, in particular, has been deemed as a main pollutant to be reduced as a result of tightening emission requirements. It has long been recognized that oxygenated alternative fuels are less polluting than diesel fuel. Dimethyl ether (DME) is one of the most attractive alternative fuels to solve the exhaust emission problems of diesel engines. The experimental studies show that diesel engines with DME fuel can achieve high thermal efficiency, ultra-low emissions, as well as softness and smokeless combustion.1–5 * To whom correspondence should be addressed. Telephone: +8613564388022. Fax: +86-21-34205553. E-mail:
[email protected]. (1) Fieisch, T.; McCarthy, C.; Basu, A.; Udovich, C.; Charbonneau, P.; Slodowske, W.; Mikkelsen, S. E.; McCandless, J. A new clean diesel technology: Demonstration of ULEV emissions on a navistar diesel engine fueled with dimethyl ether. SAE Tech. Pap. 950061, Society of Automobile Engineers, Warrendale, PA, 1995. (2) Paul, K.; Herwig, O. Development of fuel injection equipment and combustion system for DI diesels operated on dimethyl ether. SAE Tech. Pap. 950062, Society of Automobile Engineers, Warrendale, PA, 1995. (3) Hansen, J. B.; Voss, B.; Joensen, F.; Siguroardottir, I. D. Large scale manufacture of dimethyl ethersA new alternative diesel fuel from natural gas. SAE Tech. Pap. 950063, Society of Automobile Engineers, Warrendale, PA, 1995. (4) Sorenson, S. C.; Mikkelsen, S. C.; Mikkelsen, S. E. Performance and emissions of a 0.273 L direct injection diesel engine fueled with neat dimethyl ether. SAE Tech. Pap. 950064, Society of Automobile Engineers, Warrendale, PA, 1995. (5) Kajitani, S. Engine performance and exhaust characteristics of directinjection diesel engine operated with DME. SAE Tech. Pap. 972973, Society of Automobile Engineers, Warrendale, PA, 1995.
For DME engines equipped with the inline pump and using conventional in-cylinder direct-injection and combustion strategies, lower CO and hydrocarbon (HC) emissions can be obtained. Their NOx emissions are generally lower than those of diesel engines, which may meet the Europe III emission standard.6–9 However, among the above-mentioned emissions, relatively high NOx emissions were reported by several researchers.10–12 Therefore, it is very difficult for NOx emissions from DME engines with a conventional inline pump to meet a stricter standard without employing the aftertreatment. The reasons go as follows: Because of the low fuel energy and liquid density of DME, 80% more volume of DME fuel than that of (6) Hwang, J. S.; Ha, J. S.; No, S. Y. Spray characteristics of DME in conditions of common rail injection system (II). Int. J. Automot. Technol. 2003, 4, 119–124. (7) Ofner, H.; Gill, D. W.; Krotscheck, C. Dimethyl ether as fuel for CI enginessA new technology and its environmental potential. SAE Tech. Pap. 981158, Society of Automobile Engineers, Warrendale, PA, 1998. (8) Morsy, M. H.; Ahn, D. H.; Chung, S. H. Pilot injection of DME for ignition of natural gas at dual fuel engine-like conditions. Int. J. Automot. Technol. 2006, 7, 1–7. (9) Wu, J. H.; Huang, Z.; Qiao, X. Q. Study on combustion and emissions characteristics of turbocharged engine fuelled with dimethyl ether. Int. J. Automot. Technol. 2006, 7, 645–652. (10) 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, 793–800. (11) Kim, M. Y.; Yoon, S. H.; Park, K. H. Effect of multiple injection strategies on the emission characteristics of dimethyl ether (DME)-fueled compression ignition engine. Energy Fuels 2007, 21, 2673–2681. (12) Alam, M.; Fujita, O.; Ito, K.; Kajitani, S.; Oguma, M.; Machida, M. Performance of NOx catalyst in a DI diesel operated with neat dimethyl ether. SAE Tech. Pap. 1999-01-3599, Society of Automotive Engineers, Warrendale, PA, 1999.
10.1021/ef700781w CCC: $40.75 2008 American Chemical Society Published on Web 05/08/2008
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diesel fuel must be delivered in each injection. Besides, because the DME injection pressure is lower than 50 MPa under the current fuel-system technology, the duration of the DME injection is generally longer than that of diesel-fuel injection. It is unfavorable to further NOx reduction by delaying the injection timing because it can result in a higher exhaust gas temperature with a penalty of thermal efficiency.13,14 Thus, a tradeoff relation exists between NOx emissions and thermal efficiency for a diesel engine fueled with DME using a conventional in-cylinder direction-injection combustion mode. Homogeneous charge compression ignition (HCCI) combustion has been attracting growing attention in recent years because of its potentials for simultaneous reduction of exhaust gas emissions and fuel consumption in diesel engines.15–17 However, HCCI has several problems that hinder its commercialization.18 Especially, it is very difficult to control the initiation timing and extend the load range of HCCI engines. HCCI combustion with DME shows a very low NOx emissions, but CO and HC emissions turn out to be high.19–22 It can be found that a conventional in-cylinder direct-injection combustion and HCCI combustion with DME have opposite advantages and disadvantages. Taking advantage of HCCI combustion and in-cylinder directinjection combustion for a diesel engine fueled with DME, a new combustion concept, namely, compound charge compression ignition (CCCI) combustion by port aspiration and direct injection of DME, is proposed in this paper. In this concept, a portion of fuel is aspirated into the combustion chamber via the air intake port to cause HCCI combustion at the compression stroke because of its high volatility of DME, and the remainder of the fuel is injected by a conventional inline fuel pump. As a result, CCCI combustion happens, which includes HCCI combustion at first and in-cylinder spray combustion later. DME HCCI combustion shows very low NOx emission levels, and the in-cylinder injection can control the combustion and provide more engine output. The evaporation rate for fuel drops in the DME spray is much faster after HCCI combustion, which (13) Ho, T.; Gerhard, R. Fuel injection strategy for reducing NOx emissions from heavy-duty diesel engines fueled with DME. SAE Tech. Pap. 2006-01-3324, Society of Automotive Engineers, Warrendale, PA, 2006. (14) Ho, T.; James, C. M. Can heavy-duty diesel engines fueled with DME meet US 2007/2010 emissions standard with a simplified aftertreatment system. SAE Tech. Pap. 2006-01-0053, Society of Automotive Engineers, Warrendale, PA, 2006. (15) Christensen, M.; Johansson, B.; Amneus, P.; Mauss, F. Supercharged homogeneous charge compression ignition. SAE Tech. Pap. 980787, Society of Automotive Engineers, Warrendale, PA, 1998. (16) Stanglmaier, R. H.; Roberts, C. E. Homogeneous charge compression ignition (HCCI): Benefits compromises, and future engine applications. SAE Tech. Pap. 1999-01-3682, Society of Automotive Engineers, Warrendale, PA, 1999. (17) Suzuki, H.; Koike, N.; Ishii, H.; Odaka, M. Exhaust purification of diesel engines by homogeneous charge with compression ignition part 1: Experimental investigation of combustion and exhaust emission behavior under pre-mixed homogeneous charge compression ignition method. SAE Tech. Pap. 970313, Society of Automotive Engineers, Warrendale, PA, 1997. (18) Stanglmaier, R. H. Homogeneous charge compression ignition (HCCI): Benefits, compromises, and future engines applications. SAE Tech. Pap. 1999-01-3682, Society of Automotive Engineers, Warrendale, PA, 1999. (19) Li, D. G.; Huang, Z.; Qiao, X. Q. Study on HCCI combustion fueled with DME. J. Intern. Combust. Engine 2005, 23, 193–198. (20) Norimasa, L.; Tetsuya, I. Auto-ignition and combustion of n-butane and DME/air mixtures in a homogeneous charge compression ignition engine. SAE Tech. Pap. 2000-01-1832, Society of Automotive Engineers, Warrendale, PA, 2000. (21) Song, J.; Huang, Z.; Qiao, X. Q. Performance of a controllable premixed combustion engine fueled with dimethyl ether. Energy ConVers. Manage. 2004, 45, 2223–2232. (22) Zheng, Z. Q.; Yao, M. F.; Wang, Y. Experimental study of HCCI combustion process fueled with DME. J. Intern. Combust. Engine 2003, 9, 561–565.
Zhang et al. Table 1. Specifications of Proto-engine bore × stroke (mm) compression ratio rated power (kW) rated speed (r/min) plunger diameter (mm) fuel delivery advance angle (°CA BTDC) nozzle number × orifice diameter (mm) needle valve opening pressure (MPa)
135 × 145 16.5 29.4 1500 9 28 4 × 0.35 19
reduces the burning time, decreases the charge heterogeneity, and therefore, reduces NOx formation in the phase of the mixing controlled combustion. On the other hand, the gas after HCCI combustion can function as the internal exhaust gas recirculation (EGR) and dilute the charge oxygen concentration, and therefore, NOx formation is suppressed. Meanwhile, CO and HC with a relatively high level after HCCI combustion should be further oxidated during the period of the in-cylinder spray combustion. Because only partial demand needs to be injected and DME evaporates almost immediately, combustion duration for the direct injection plus the unburnt fuel in the cylinder is not long. In this concept, the ratio of DME of port aspiration to in-cylinder injection and the injection timing can be regulated according to the engine load. Furthermore, the port fuel design concept and external EGR can be implemented to control HCCI combustion phasing, keep high thermal efficiency, and avoid an excessive rate of pressure rise, which causes knocking. It is believed that CCCI combustion can extend the operating range with little change of NOx emissions and a dramatic reduction of HC and CO emissions in comparison to the HCCI combustion mode. Using the methods of port fuel design and oxidizer improvement to control the ignition timing of HCCI combustion, DME fuel consumption of the CCCI engine could be improved. In this study, we attempted to improve combustion and reduce emissions by means of direct-injection (DI) fuel injection timing control and regulating the ratio of port aspirated DME to injected DME in the cylinder. To control the ignition and combustion phase of HCCI engines, reduce engine knock, and expand the engine load range, DME of intake port aspiration was mixed with liquid petroleum gas (LPG), which has a good antiknock property and the effect of LPG percentage in DME/LPG blended fuel on combustion characteristics was also investigated. To further hinder NOx emissions, port introduction of CO2 gas was employed to function as EGR and the effect of EGR on CCCI combustion and emission characteristics was evaluated. 2. Experimental Apparatus A two-cylinder, four-stroke naturally aspirated high-speed DI diesel engine was modified to conduct the CCCI combustion test. The specifications of the prototype engine and the schematic of the DME CCCI combustion experimental system are shown in Table 1 and Figure 1, respectively. A vaporizer was fixed in the fuel supply system to produce DME vapor, and a mixer was installed in front of the intake port where an air/fuel homogeneous mixture was formed. The mass flow can be controlled by a regulator. As for the in-cylinder direct-injection fuel supply system, a lowpressure pump was installed in front of the fuel injection pump to increase the inlet pressure of the injection pump to prevent vapor lock in the fuel injection system. Because the lower heating value of liquid DME is only 2/3 that of diesel fuel, to deliver the same fuel energy as in 1 kg of diesel fuel (the required DME is 1.49 kg), a larger amount of fuel supply per cycle is needed to ensure the same engine power. Therefore, nozzle number × orifice diameter changed from 4 × 0.35 to 5 × 0.43 mm, and needle valve opening pressure decreased to 15 MPa. A pressure regulator and a buffer were employed to maintain the fuel supply pressure and
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Figure 1. Schematic of the DME CCCI experimental system. Table 2. Physical and Chemical Properties of DME and LPG chemical formula mole weight (g) boiling point (°C) Reid vapor pressure (MPa) liquid density (g/cm3) low heat value (MJ/kg) ignition temperature (°C) cetane number stoichiometric air/fuel ratio (kg/kg) latent heat of evaporation (kJ/kg) percent weight of carbon percent weight of hydrogen percent weight of oxygen
DME
LPG
CH3-O-CH3 46.07 -24.9 0.51 (20 °C) 0.668 27.6 235 55-60 9.0 460 (-20 °C) 52.2 13.0 34.8
C3H8, C4H10 44-56 -42.1 0.84 (20 °C) 0.501 49.79 470 3 15.58 426, 385 81.8 18.2, 17.9 0
eliminate the pressure pulsation. About 2% castor oil was added in DME fuel as a lubricant for the combustion study. To evaluate the effect of EGR on CCCI combustion, port introduction of CO2 gas was employed to function as EGR. As shown in Figure 1, a CO2 tank was connected to the air intake port by a pipe and the flow mass of CO2 can be measured by a flow meter when the CO2 gas with a high pressure was relieved by a pressure relief valve. The relief valve was used to control the amount of mass of CO2. The mass flow of fresh air charge was measured by a laminar flow meter installed in the air inlet pipe, as shown in Figure 1.
3. Experimental Procedure To investigate the effect of the premixed charge on exhaust emissions and combustion characteristics, the experiment was conducted using the experimental system shown in Figure 1, which can control the premixed fuel and directly injected DME. Specifications of DME and LPG used in this work are listed in Table 2. The premixed ration rp for neat DME and the DME/LPG-blended fuels can be obtained according to the following equation: rp )
mp_DMEhDME + mp_LPGhLPG (mp_DME + md_DME)hDME + mp_LPGhLPG
(1)
mp_DME is the mass of DME through air intake port, mp_LPG is the mass of LPG in the blended fuels, and md_DME is the mass of directly injected DME. hDME and hLPG are the lower heat values of DME and LPG, respectively. A different premixed ratio can be achieved by regulating the DME mass of port aspiration and in-cylinder injection. From the above equation, it is found that mere HCCI combustion occurs when rp is equal to 1 and mere direct-injection combustion occurs when rp is equal to 0. In this work, the mass flow of DME by air intake port and in-cylinder-injected DME was measured by two electronic scales, respectively. CO2 concentration in air charge can be obtained by the following equation: CO2 (%) )
CO2,intake × 100% CO2,intake + airintake
(2)
CO2,intake and airintake are the mass flows aspirated into the combustion chamber. The CO2 concentration in air charge can be regulated by the pressure relief valve. The engine speed was 1000 r/min in this test, held to within (2 rpm. To guarantee the repeatability and comparability of the measurements for different operating conditions, the intake charge temperature was fixed at 10 °C, held to within (2 °C. The coolantout temperature remained at 85 °C, held to within ( 2 °C. At the engine speed of 1000 r/min, φc (charge efficiency) is about 0.828 at the indicated mean effective pressure (IMEP) of 0.35 MPa and about 0.817 at the IMEP of 0.525 MPa. In this work, there was a little fluctuation for the intake pressure during the progress of recording, which was about 0.0975-0.0983 MPa for all runs. The cylinder pressure was measured with a Kistler Model 6125A pressure transducer. The charge output from this transducer was converted to an amplified voltage with a Kistler Model 5015 amplifier. The 1440 pulses per rotation (4 pulses per crank angle) from a shaft encoder on the engine crankshaft were used as the data acquisition clocking pluses to acquire the cylinder pressure data. Pressure data were recorded using high-speed memory. For each measuring point, the pressure data of 40 consecutive cycles were sampled and recorded. The pressure trace for a specific condition was obtained by averaging the sampled pressure data. CO, HC, and NOx emissions were measured by an AVL Digas 4000 gas analyzer. On the basis of the measured combustion pressure, the rate of heat release (ROHR) and the mean gas temperature were calculated using the traditional first law heat-release model.
4. Results and Discussion 4.1. Study on Combustion and Emissions for Neat DME. 4.1.1. Study on Combustion Characteristics. Figure 2 shows the effect of the premixed fuel ratio on in-cylinder mean temperature, pressure, and ROHR in relation to various injection timing at the IMEP of 0.35 MPa. At rp ) 0.0, a conventional direct-injection combustion was presented, which includes a premixing combustion and a diffusion combustion. After port aspiration of DME, at rp ) 0.19, a cool-flame stage, which features HCCI combustion, was observed. However, a thermal flame stage was not obvious because of a lean fuel/air mixture. After HCCI combustion, a premixing combustion and a diffusion combustion by injected fuel were presented. As shown in Figure 2, the thermal flame reaction of HCCI combustion may overlap with the in-cylinder spray diffusion combustion by in-cylinder-injected DME fuel and the lean mixture of DME/air should be completely burned after in-cylinder spray diffusion combustion. Therefore, the overall combustion process at rp ) 0.19 comprised three stages, and the peak values of ROHR, in-cylinder mean temperature, and pressure were lower than rp ) 0.0. Because of HCCI combustion, the pressure and temperature of gases in the combustion chamber increased, thus reducing the ignition delay of the injected DME fuel. Hence, ROHR in the early combustion stage for injected DME fuel was lower. At rp ) 0.31, the peak value of ROHR for a cool-flame stage increased slightly because of a richer mixture compared to that of rp ) 0.19 and the peak values of ROHR for the premixing combustion, in-cylinder mean temperature, and pressure decreased. At rp ) 0.52, HCCI combustion of DME typically exhibited a two-stage combustion characteristic and a region called the negative temperature coefficient (NTC) was obviously displayed, while the ignition timing of CCCI combustion started in advance. The combustion of injected DME fuel only presented a diffusion combustion mode. The overall combustion process also displayed a threestage combustion with a different characteristic compared to that of rp ) 0.19 and 0.31.
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Figure 2. In-cylinder mean temperature, pressure, and ROHR in relation to various DI fuel injection timing for (a) rp ) 0.0, (b) rp ) 0.19, (c) rp ) 0.31, and (d) rp ) 0.52 (IMEP ) 0.35 MPa, n ) 1000 r/min).
It needs to be pointed out that, with the HCCI combustion method, the maximum IMEP in this work was 0.20 MPa because of a knock limit and the maximum premixed fuel ratio was rp ) 0.70 at 0.35 MPa IMEP. It is found that, under the same premixed fuel ratio, the peak values of in-cylinder mean temperature and pressure increased with the advance of DI fuel injection timing and the beginning of combustion for in-cylinder-injected fuel advanced. However, the change of ignition timing of HCCI combustion was not obvious. Figure 3 shows the effect of the premixed fuel ratio on the pressure rise rate at the fuel delivery advance angle of 22°CA BTDC at 0.35 MPa IMEP. In this case, the peak value of the pressure rise rate decreased at first and increased later. The reason for this decrease was that HCCI combustion with lean fuel/air mixtures raised the charge temperature and shortened the ignition delay of directinjected fuel. However, rich fuel/air mixtures resulted in a fast high-temperature reaction of HCCI combustion, and the charge
temperature increased with a high degree prior to the event of in-cylinder direct injection. The peak value of the pressure rise rate for rp ) 0.65 was higher than that of in-cylinder injection. The pressure rise rate can display the speed for combustion and ROHR. The faster the pressure rise rate, the more pressure oscillations, thus affecting the stability of engine work. Therefore, using an appropriate premixed fuel ratio, the peak value of the pressure rise rate can be controlled to ensure a stable engine operation. 4.1.2. Study on Emissions and Fuel Consumption. Figure 4 shows the effect of the premixed fuel ratio on fuel consumption, emissions, and indicated thermal efficiency for various DI fuel timing injections at 0.35 MPa IMEP. In Figure 4, bi means the indicated fuel consumption, which is defined by B × 1000/Pi, where B (kg/h) is the DME fuel consumption per hour and Pi (kw) is the indicated power calculated by the cylinder pressure. The indicated thermal efficiency ηi can be obtained by the following equation:
Optimizing Control Method of CCCI Combustion
ηi )
3.6 × 106 hDMEbi
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(3)
hDME is the lower heat values of DME, which is 27 600 kJ/kg, as can be found in Table 2. NOx is a combination of nitrogen monoxide (NO), nitrogen dioxide (NO2), and nitrous oxide (N2O). The formation of oxides of nitrogen is a side effect of combustion. NOx does not have a strong correlation to the fuel structure but greatly depends upon how the combustion is organized. In CCCI engines, NOx primarily forms outside sprays of directly injected fuel, where the temperature is higher than that of inside sprays and where there are less hydrocarbons competing with nitrogen for oxygen. From Figure 4a, it can be seen that NOx emissions decreased at first and increased later with the rise of the premixed fuel ratio for various DI fuel timing injections. HCCI combustion by a lean fuel/air mixture via air intake port slightly increased the charge temperature and pressure, which shortened the ignition delay of injected DME fuel, reduced the ROHR in the premixed combustion phase, and thus suppressed the NOx formation. Meanwhile, the combustion products after HCCI combustion diluted the oxygen concentration in the charge, which reduced the driving force for NOx formation. The reason for the increase of NOx emissions hereafter was that a richer fuel/air mixture resulted in a rapid high-temperature reaction, which increased the charge temperature at the start of combustion (SOC) to a higher level as can be seen in Figure 2. A higher temperature increased the driving force for NOx formation. However, using an appropriate premixed fuel ratio and fuel delivery advance angle, NOx emissions can be reduced to near-zero levels. It is found that, under the same fuel delivery advance angle, HC emissions increased with the port aspiration of DME, which was mainly due to the incompletely burned DME fuel between the piston and cylinder wall. With the increase of the premixed fuel ratio, the maximum of HC emissions was maintained at around 260 ppm. It is interesting to note that CO emissions presented the trend of increasing at first and decreasing later with the increase of the premixed fuel ratio. The reason for the increase of CO emissions was that, with the port aspiration of DME, HCCI combustion of a lean fuel/air mixture decreased the in-cylinder gas temperature, as can be found in Figure 2. The reason for the decrease in CO emissions was that, with a high premixed fuel ratio, HCCI combustion increased the in-cylinder gas Figure 4. (a) Fuel consumption and emissions and (b) indicated thermal efficiency versus premixed fuel ratios for various DI fuel timing injections (IMEP ) 0.35 MPa, n ) 1000 r/min).
Figure 3. Effect of the premixed fuel ratio on the pressure rise rate.
temperature considerably and ignition timing of CCCI combustion advanced further, both of which were very beneficial to the oxidation of CO. As can be seen from Figure 4a, at the same premixed fuel ratio, the indicated fuel consumption was improved with the increase of the fuel delivery advance angle and CO and HC emissions decreased. This was mainly due to the increase in the gas temperature under an early injection timing. HCCI combustion occurred with the port aspiration of DME, and the fuel near compression clearances and the cylinder wall was hard to be burned completely. The higher the premixed fuel ratio, the more unburned DME fuel. Meanwhile, improper combustion phasing of neat DME HCCI combustion led to a poor combustion efficiency. Hence, to improve DME fuel consumption of the CCCI engine, one of the main thrusts of this work is to gain control of the ignition timing of HCCI combustion by port fuel design or oxidizer improvement, etc.
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Figure 5. In-cylinder mean temperature, pressure, and ROHR with DME/LPG blended fuel in fresh air at (a) 0.35 MPa IMEP and (b) 0.525 MPa IMEP (n ) 1000 r/min, fuel delivery advance angle ) 22°CA BTDC).
As shown in Figure 4b, with the increase of the premixed fuel ratio, the indicated thermal efficiency decreased under the same injection timing. It was found from eq 3 that, compared to the indicated fuel consumption, the indicated thermal efficiency can play the same role to illustrate the energy conversion efficiency for CCCI engine fueled with DME fuel. 4.2. Study on Combustion and Emissions for DME/LPG Blended Fuel by Air Intake Port. 4.2.1. Study on Combustion Characteristics. Figure 5 shows the in-cylinder mean temperature, pressure, and ROHR for CCCI combustion with the blended fuels under different engine loads. LPG has a similar physical property to DME, which has a low saturation pressure
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and is easy to vaporize. However, the LPG has a lower cetane number and a good antiknock property. Both LPG and DME have good solubility. When the ratio of LPG to DME in the blended fuels is changed, it is practical to obtain a different cetane number to control the ignition and combustion phase of HCCI, thus improving the fuel consumption and emissions of the CCCI engine. In this work, the CCCI engine performance with a LPG percentage of 15 and 41% in DME/LPG blended fuel was investigated. To make a compromise between DME fuel consumption and emissions, the fuel delivery advance angle with 22°CA BTDC was selected. In Figure 5, “DME/HCCI” means that neat DME is aspirated by the port before in-cylinder direct injection to conduct HCCI combustion and “59% DME and 41% LPG/HCCI” means that the blended fuel with the DME percentage of 59% and LPG of 41% is aspirated by the port before in-cylinder direct injection to conduct HCCI combustion. As can be seen from Figure 5a that, for the premixed fuel ratio of rp ) 0.68 at 0.35 MPa IMEP, peak values of in-cylinder gas temperature and pressure were lower with the blended fuel than the neat DME mainly because of a lower cetane number for the blended fuel and the delayed starting time of HCCI combustion. However, the peak value of ROHR for a thermal flame stage slightly increased for the blended fuel because the phase of the thermal flame stage was relatively close to the top death center (TDC), which was very beneficial to the improvement of the fuel economy. It is found from Figure 5b that, for the premixed fuel ratio of rp ) 0.54 at 0.525 MPa IMEP, the peak values of the incylinder temperature and pressure decreased with the increase in LPG percentage. Meanwhile, the starting time for HCCI combustion was gradually delayed, and the centroid of ROHR for CCCI combustion was moving toward TDC, which should lead to a good thermal efficiency. 4.2.2. Study on Emissions and Fuel Consumption. Figure 6 shows the effect of the premixed fuel ratio on fuel consumption and emissions at different engine loads with the DME/LPG blended fuel in the fresh charge. With the same load, it is found that NOx emissions decreased with the increase of the LPG percentage in blended fuel. The ignition timing for HCCI combustion was delayed, and the incylinder gas temperature was lower than that of neat DME prior to the combustion of in-cylinder injection, which resulted in a lower level of NOx emissions. Meanwhile, a lower in-cylinder gas temperature after HCCI combustion with the blended fuel led to an increase of CO and HC emissions. It is found that, at 0.35 MPa IMEP with the blended fuel, DME fuel consumption for the CCCI DME engine was improved in a high premixed fuel ratio. At 0.525 MPa IMEP, fuel consumption presented an improvement for all premixed fuel ratios. The higher the percentage in the blended fuel, the lower the fuel consumption. As for the same LPG percentage in the blended fuel, the emissions and fuel consumption for CCCI combustion presented similar behaviors to that of the neat DME. However, with the LPG percentage of 41%, when the CCCI engine ran at 0.35 MPa IMEP, the internal EGR after HCCI combustion played an important role in suppressing the formation of NOx emissions. Therefore, the NOx emissions decreased monotonously with the increase of the premixed fuel ratio. 4.3. Evaluation of the Effect of CO2 on CCCI Combustion in an Engine Fueled with Dimethyl Ether. 4.3.1. Study on Combustion Characteristics. Figure 7 shows the effect of the CO2 concentration in air charge on in-cylinder mean temperature, pressure, and ROHR.
Optimizing Control Method of CCCI Combustion
Figure 6. Fuel consumption and emissions versus premixed fuel ratios with DME/LPG blended fuel at (a) 0.35 MPa IMEP and (b) 0.525 MPa IMEP (n ) 1000 r/min, fuel delivery advance angle ) 22°CA BTDC).
Figure 7a shows that, for a premixed fuel ratio of rp ) 0.25, a cool-flame stage for HCCI combustion was observed and the thermal flame stage was not obvious. The combustion of injected fuel exhibited a two-stage combustion, including premixing combustion and diffusion combustion. The overall combustion process displayed three stages. With the port aspiration of CO2 gas, the starting time of the cool-flame stage and injected fuel combustion were both delayed and the peak value of ROHR for the premixing combustion of the injected fuel decreased. With 13% CO2 concentration in air charge, the in-cylinderinjected fuel combustion almost showed a single peak combustion characteristic, which is similar to a modulated kinetics
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Figure 7. In-cylinder mean temperature, pressure, and ROHR with a variation of the CO2 concentration in fresh air at (a) rp ) 0.25 and (b) rp ) 0.40 (IMEP ) 0.35 MPa, n ) 1000 r/min, fuel delivery advance angle ) 14°CA BTDC).
(MK).23,24 As a result, HCCI-MK combustion for CCCI combustion engines could be observed. It should be pointed out that, because the fuel delivery advance angle of 14°CA BTDC is too late for the DME-fueled engine with the conventional injection combustion method, the (23) Matsui, Y. A new combustion concept for small DI diesel enginess1st report: Introduction of the basic technology. JSAE9730416, 1997. (24) Kimura. S. A new combustion concept for small DI diesel enginess2nd Report: Effects on engine performance. JSAE9735051, 1997.
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0.40. HC emissions were high with a low premixed fuel ratio. This was mainly due to the incompletely burned DME fuel under a relatively late injection timing. With the increase of the premixed fuel ratio, the in-cylinder mean temperature increased and HC emissions decreased gradually. CO emissions presented the characteristic of increasing at first and decreasing later, which was similar to the case of the neat DME. With the port aspiration of CO2 gas, NOx emissions were almost zero. However, CO and HC emissions as well as fuel consumption increased slightly. 5. Conclusions
Figure 8. Comparison of fuel consumption and emissions with a variation of CO2 in fresh air (IMEP ) 0.35 MPa, n ) 1000 r/min, fuel delivery advance angle ) 14°CA BTDC).
misfire for engine operation happens. Under the help of HCCI combustion, the engine can run stably. As can be seen from Figure 7b, at rp ) 0.40, HCCI combustion of DME exhibited a two-stage combustion characteristic and the injected fuel presented a conventional combustion method including the premixing combustion and a diffusion combustion. The overall combustion process comprised four stages. With the increase of the CO2 concentration, the starting time of CCCI combustion was delayed and the proportion of the premixing combustion of the injected fuel increased. HCCI-MK combustion could not be observed becaus HCCI combustion with a rich fuel/air mixture led to a higher in-cylinder temperature prior to the in-cylinder injection. Hence, the injected fuel showed a conventional in-cylinder directinjection combustion. 4.3.2. Study on Emissions and Fuel Consumption. Figure 8 shows the effect of the CO2 concentration in air charge on fuel consumption and emissions. It can be seen that NOx emissions were lower than 6 ppm with neat DME and CCCI combustion presented almost zero NOx emissions when the premixed fuel ratio was larger than
(1) CCCI combustion process comprises HCCI combustion, a premixing combustion, and a diffusion combustion. The combustion characteristic is mainly determined by the premixed fuel ratio. (2) As for the neat DME fuel, with the port aspiration of DME, HC emissions increased. With an increase of the premixed fuel ratio, CO emissions increased first but decreased later and NOx emissions decreased first but increased later. Meanwhile, DME fuel consumption suffered from improper combustion phasing. (3) As for the same premixed fuel ratio, with the advance of the injection timing, peak values of incylinder temperature and pressure increased and the beginning of combustion of in-cylinder-injected fuel advanced. NOx emissions increased, but HC and CO emissions decreased. As a result, the thermal efficiency was improved. (4) After port aspiration of the DME/LPG-blended fuel, at the same load, peak values of in-cylinder temperature and pressure decreased gradually when compared to neat DME. (5) With the increase of the LPG percentage in the blended fuel, NOx emissions decreased and thermal efficiency was improved at 0.35 MPa IMEP with a high premixed fuel ratio. Fuel consumption was decreased at 0.525 MPa IMEP for all premixed ratios. (6) With an appropriate CO2 concentration in air charge, the HCCI-MK combustion concept for CCCI combustion engines was observed and NOx emissions can be lowered to near-zero levels with this combustion method. Acknowledgment. This work is supported by the National Natural Science Foundation of China, 50276035, and a specialized Research Fund for the Doctoral Program, The Ministry of Education, 20050248013. EF700781W
