1498 Energy & Fuels 2006, 20, 1498-1504 Combustion Characteristics of a Direct-Injection Engine Fueled with Natural Gas-Hydrogen Blends under Various Injection Timings Zuohua Huang,* Jinhua Wang, Bing Liu, Ke Zeng, Jinrong Yu, and Deming Jiang State Key Laboratory of Multiphase Flow in Power Engineering, Xi'an Jiaotong UniVersity, Xi'an 710049, P. R. China ReceiVed January 23, 2006. ReVised Manuscript ReceiVed April 21, 2006 Combustion characteristics under various injection timings of a direct-injection engine fueled with natural gas-hydrogen blends at fixed injection duration and fixed ignition timing were investigated. The study shows that early injection decreases the excessive-air ratio and makes leaner mixtures. The brake mean effective pressure increases with the advancement of fuel-injection timings. The brake mean effective pressure reaches a maximum value at an injection timing of 190 CA BTDC and maintains this maximum value with the further advancement of fuel-injection timings. For a specific injection timing, an increase in the hydrogen fraction decreases the brake mean effective pressure when the hydrogen fraction is less than 10%, whereas the brake mean effective pressure tends to increase when the hydrogen fraction is larger than 10%. Combustion durations decrease with the advancement of fuel-injection timing. When the hydrogen fraction is less than 10%, combustion durations increase with increasing hydrogen fractions; conversely, combustion durations decrease with increasing hydrogen fractions when the hydrogen fraction is larger than 10%. The amounts of NO x and CO 2 increase with advancing fuel-injection timing, and the CO concentration experiences small variations under various fuel-injection timings. The addition of hydrogen in natural gas can reduce the CO 2 concentration. Introduction With increasing concerns about energy shortages and environmental protection, research on improving engine fuel economy and reducing exhaust emissions has become a major research focus in combustion and engine development. Because of limited reserves of crude oil, the development of alternative fuel engines has attracted more and more concern in the engine community. Alternative fuels usually belong to the category of clean fuels, rather than diesel and gasoline fuel, in the engine combustion process. The introduction of these alternative fuels is beneficial for slowing down the fuel shortage and reducing engine exhaust emissions. Natural gas is considered to be one of the favorable fuels for engines, and the natural-gas-fueled engine has been realized in both the spark-ignited engine and the compressionignited engine. However, because of the slow-burning velocity of natural gas and its poor lean-burn capability, the natural gas spark-ignited engine has the disadvantage of large cycle-bycycle variations, which will decrease the engine power output and increase fuel consumption. 1-2 Because of these restrictions, a natural gas engine is usually operated under the condition of a stoichiometric equivalence ratio with relatively low thermal efficiency. The traditional homogeneous spark-ignited natural gas engine has the disadvantage of low volumetric efficiency, because natural gas occupies a fraction of the intake charge; this * Corresponding author. Address: School of Energy and Power Engineering, Xi an Jiaotong University, Xi an 710049, P. R. China. E-mail: zhhuang@mail.xjtu.edu.cn. (1) Rousseau, S.; Lemoult, B.; Tazerout, M. Combustion characteristics of natural gas in a lean burn spark-ignition engine. IME Proc., Part D: J. Automobile Eng. 1999, 213 (D5), 481-489. (2) Ben, L.; Dacros, N. R.; Truquet, R.; Charnay, G. Influence of Air/ Fuel Ratio on Cyclic Variation and Exhaust Emission in Natural Gas SI Engine. SAE Technical Paper no 992901; Society of Automotive Engineers: Warrendale, PA, 1999. decreases the amount of fresh air going into the cylinder and reduces the power output compared to that of a gasoline engine. Meanwhile, the homogeneous charge combustion makes it difficult to burn the lean mixture. The introduction of natural gas direct-injection combustion can help avoid the loss in volumetric efficiency, as natural gas is directly injected into the cylinder and is flexible in mixture preparation; for example, it may form a stratified mixture in the cylinder at lean mixture combustion, improving the fuel economy to be on par with that of the direct-injection gasoline engine. 3-4 Traditionally, to improve the lean-burn capability and flameburning velocity of the natural gas engine under lean-burn conditions, we introduce an increase in the flow intensity in the cylinder; this measure always increases the heat loss to the cylinder wall and increases the combustion temperature as well as the NO x emission. 5 One effective method for solving the problem of the slow-burning velocity of natural gas is to mix the natural gas with a fuel that possesses a fast-burning velocity. Hydrogen is regarded as the best gaseous candidate for natural gas because of its very fast burning velocity; this combination is expected to improve the lean-burn characteristics and decrease engine emissions. 4 Blarigan and Keller investigated the port-injection engine fueled with natural gas-hydrogen mixtures; 6 Akansu and Dulger studied engine fueld with natural gas-hydrogen mixtures. 7 Wong and Karim studied engine performance fueled by various (3) Zeng, K.; Huang, Z. H.; Liu, B.; Liu, L. X.; Jiang, D. M.; Ren, Y.; Wang, J. H. Combustion Characteristics of a Direct-Injection Natural Gas Engine under Various Injection Timings. Appl. Therm. Eng. 2006, 26(8-9), 806-813. (4) Huang, Z. H.; Wang, J. H.; Liu, B.; Zeng, K.; Yu, J. R.; Jiang, D. M. Combustion characteristics of a direct-injection engine fueled with natural gas-hydrogen mixtures. Energy Fuels 2006, 20 (2), 540-546. (5) Das, A.; Watson, H. C. Development of a Natural Gas Spark-Ignition Engine for Optimum Performance. IME Proc., Part D: J. Automobile Eng. 1997, 211 (D5), 361-378. 10.1021/ef060032t CCC: $33.50 2006 American Chemical Society Published on Web 05/20/2006
Combustion Characteristics at Various Injection Times Energy & Fuels, Vol. 20, No. 4, 2006 1499 hydrogen fractions in natural gas-hydrogen blends, 8 and Bauer and Forest conducted an experimental study on natural gashydrogen combustion in a CFR engine. 9 Furthermore, studies on the lean-combustion capability of natural gas-hydrogen combustion and natural gas-hydrogen combustion with turbochargingand/orexhaustgasrecirculationwerealsoconducted. 10-12 These studies showed that the exhaust HC, CO, and CO 2 concentrations of an engine operating on natural gas-hydrogen blends could be lower than the exhaust concentrations from a natural gas engine. However, NO x may increase in natural gashydrogen combustion at rich mixture conditions as the leanburn ability improves and the flame-propagation speed increases; the NO x concentration can be greatly decreased through lean combustion and the retarding of the ignition advance angle. The previous work mainly concentrated on a homogeneous mixture fueled from the port, and few studies have been reported on a direct-injection engine. Shudo and Shimamura investigated the combustion and emissions of an engine with port-injected hydrogen and in-cylinder injection natural gas. 13 This type of engine needs two separate fueling systems, which make the system complicated. Huang and Shiga investigated the influence of fuel-injection timings on direct-injection natural gas combustion using a compression ignition machine and revealed the importance of injection timing on combustion and emissions. 14-15 Because the fuel-injection timing is very important to the optimization of the direct-injection gas engine, and because no previous work has been done in this aspect, this study will concentrate on the investigation of the effects of fuel-injection timings on combustion characteristics of an engine fueled with direct-injection natural gas-hydrogen mixtures so as to provide practical guidance for engine optimization. item Table 1. Engine Specifications bore (mm) 100 stroke (mm) 115 displacement (cm 3 ) 903 compression ratio 8 combustion chamber bowl-in-shape injection pressure (MPa) 8 ignition source spark plug Table 2. Compositions of Natural Gas a fraction (vol %) item fraction (vol %) items fraction (vol %) CH 4 96.16 C 2H 6 1.096 C 3H 8 0.136 ic 4H 10 0.021 nc 4H 10 0.021 ic 5H 12 0.006 nc 5H 12 0.005 N 2 0.001 CO 2 2.54 H 2S 0.0002 H 2O 0.006 a Volumetric higher heating value: 36.588 MJ/m 3 (normal temperature and pressure).volumetric lower heating value: 32.970 MJ/m 3 (normal temperature and pressure) Table 3. Fuel Properties of Natural Gas and Hydrogen fuel property natural gas hydrogen density in 1 atm, 300 K (kg/m 3 ) 0.754 0.082 stoichiometric air-to-fuel ratio (vol %) 9.396 2.387 stoichiometric air-to-fuel ratio (wt %) 0.062 0.029 laminar flame speed (m/s) 0.38 2.9 quenching distance (mm) 1.9 0.6 mass lower heating value (MJ/kg) 43.726 119.7 volumetric heating value (MJ/m 3 ) 32.97 10.22 octane number 120 C:H ratio 0.2514 0 Experimental Section A single cylinder engine was modified into a natural gas directinjection engine. The specifications of the engine are listed in Table 1. The injector used in the study is modified from a gasoline directinjection engine made by thitachi Co. To increase the flow rate for natural gas application, we removed the swirler near the tip of (6) Blarigan, P. V.; Keller, J. O. A Hydrogen-Fueled Internal Combustion Engine Designed for Single Speed/Power Pperation. Int. J. Hydrogen Energy 2002, 23 (7), 603-609. (7) Akansu, S. O.; Dulger, A.; Kahraman, N. Internal Combustion Engines Fueled by Natural Gas-hydrogen Mixtures. Int. J. Hydrogen Energy 2004, 29 (14), 1527-1539. (8) Wong, Y. K.; Karim, G. A. An Analytical Examination of the Effects of Hydrogen Addition on Cyclic Variations in Homogeneously Charged Compression-Ignition Engines. Int. J. Hydrogen Energy 2000, 25 (12), 1217-1224. (9) Bauer C. G.; Forest, T. W. Effect of Hydrogen Addition on the Performance of Methane-Fueled Vehicles. Part I: Effect on S.I. Engine Performance. Int. J. Hydrogen Energy 2001, 26 (1), 55-70. (10) Sierens, R.; Rosseel, E. Variable Composition Hydrogen/Natural Gas Mixtures for Increased Engine Efficiency and Decreased Emissions. J. Eng. Gas Turbines Power 2000, 122 (1), 135-140 (11) Larsen, J. F.; Wallace, J. S. Comparison of Emissions and Efficiency of a Turbocharged Lean-Burn Natural Gas and Hythane-Fueled Engine. J. Eng. Gas Turbines Power 1997, 218 (1), 218-226. (12) Allenby, S.; Chang, W. K.; Megaritis, A.; Wyszynski, M. L. Hydrogen Enrichment: A Way To Maintain Combustion Stability in a Natural-Gas-Fueled Engine with Exhaust-Gas Recirculation, the Potential of Fuel Reforming. IME Proc., Part D: J. Automobile Eng. 2001, 215 (D3), 405-418. (13) Shudo, T.; Shimamura, K.; Nakajima, Y. Combustion and Emissions in a Methane DI Stratified Charge Engine with Hydrogen Premixing. JSAE ReV. 2000, 21 (1), 3-7. (14) Z. Huang, Z.; Shiga, S.; Ueda, T. Effect of Fuel-injection Timing Relative to Ignition Timing on Natural Gas Direct-Injection Combustion. J. Eng. Gas Turbines Power 2003, 125 (3), 465-475. (15) Huang, Z.; Shiga, S.; Ueda, T. Combustion Characteristics of CNG Direct-Injection Combustion under Various Fuel-Injection Timing. IME Proc., Part D: J. Automobile Eng. 2003, 217 (D5), 393-401. Figure 1. Lower heating value of fuel blends versus hydrogen fractions. the nozzle. The calibration of the pulse width with the injection amount was made by the manufacturer as well as by the authors. The flow rate of the injector under 9 MPa was 193 L/min. In addition to installing the natural gas high-pressure injector, we also installed a spark plug at the center of the combustion chamber as the ignition source. Natural gas was injected into the cylinder at a constant pressure of 8 MPa; the gas velocity from the injector nozzle is kept at a constant value of sonic velocity because of the condition of choke flow during the fuel injection, and thus the amount of injected fuel will stay at a constant value determined by the injection duration in this study. Hydrogen with 99.995% purity was used, and natural gas constitutions are given in Table 2. The fuel properties of natural gas and hydrogen are given in Table 3. Different fractions of natural gas-hydrogen mixtures were prepared in advance in a fuel bomb and were supplied to the fuel injector. Sonic flow of the injected gases was presented because of the choke flow during injection. It is estimated that an 18% volume fraction of hydrogen corresponds to a 2% mass fraction of hydrogen in the mixture, and thus the influence on the volumetric flow rate is limited. Therefore, the volumetric flow rate of natural gas-hydrogen mixtures in this study is assumed to be unchangeable and can be regarded as a function of injection duration. Figure 1 gives the volumetric heat value of natural gashydrogen-air mixtures versus the hydrogen fractions at the
1500 Energy & Fuels, Vol. 20, No. 4, 2006 Huang et al. Figure 2. C:H ratio of fuel blends versus hydrogen fractions. Figure 4. Fuel-injection and fuel-ignition timing settings under various injection timings. Figure 3. Heating value of natural gas-hydrogen-air mixtures versus excessive-air ratios. stoichiometric equivalence ratio. It can be seen that the volumetric heat value of natural gas-hydrogen mixtures will decrease with an increase in hydrogen fraction in the fuel blends, which is attributable to the low volumetric heat value of the hydrogen-air mixture compared to that of the natural gas-air mixture at the stoichiometric equivalence ratio condition. Thus, for a given fuelinjection duration, the amount of heat released will decrease with an increase in hydrogen fraction in the fuel blends. To maintain the same equivalence ratio, we must inject more fuel for the natural gas-hydrogen mixture combustion. An 8% reduction in the volumetric heating value is presented when 10% hydrogen in volume is added in natural gas. Figure 2 shows the hydrogen:carbon (H:C) ratio of natural gashydrogen blends versus the hydrogen fractions. The H:C ratio increases linearly with an increase in hydrogen fraction in the fuel blends, which will be beneficial to the reduction of carbon-related emissions such as HC, CO, and CO 2. Figure 3 gives the volumetric heating value of natural gashydrogen blends versus the excessive-air ratio. The figure shows that the volumetric heating value decreases linearly with an increase in the excessive-air ratio. The same decreasing rate-of-heating value to excessive-air ratio is presented for both natural gas and natural gas-hydrogen mixtures. Four types of fuels were prepared for the experiment: pure natural gas, a fuel blend with 85% natural gas and 15% hydrogen in volume, a fuel blend with 90% natural gas and 10% hydrogen in volume, and a fuel blend with 82% natural gas and 18% hydrogen. In this study, six fuel-injection timings were selected from the range between 160 and 210 CA BTDC under a fixed engine speed (1200 r/min), fixed fuel-injection duration (16.5 ms, corresponding to 119 CA), fixed ignition advance angle (32.5 CA BTDC), and fixed throttle opening (70% of WOT). Thus, the influence of fuel-injection timings on direct-injection natural gashydrogen can be reflected by varying the injection timings and maintaining other parameters at fixed values. Figure 4 shows the setting of fuel-injection timings in this study. Injection timing starts before the intake valve closes (φ ivc ), and part of the fuel will be injected into the cylinder before the intake valve closes. This will increase the cylinder pressure and reduce the amount of fresh air flowing into the cylinder, which will decrease the excessive-air ratio of the cylinder charge. Early injection will decrease the excessive-air ratio and make leaner mixtures. Instrumentation and Method of Calculation. The cylinder pressure was recorded by a piezoelectric transducer with the resolution of 10 Pa, and the dynamic top-dead-center (TDC) was determined by motoring. The crank angle signal was obtained from an angle-generating device mounted on the main shaft. The signal of the cylinder pressure was acquired for every 0.5 CA, and the acquisition process covered 254 completed cycles; the average value of these 254 cycles was outputted as the pressure data used for the calculation of combustion parameters. A thermodynamic model is used to calculate the thermodynamic parameters in this paper. The model neglects the leakage through the piston rings, 14 and thus the energy conservation in the cylinder is written as follows dφ - dq W dφ ) d(mu) dφ The gas-state equation is The variation in the gas-state equation with crank angle is given by The heat-release rate ( /dφ) can be derived from eqs 1 and 3 as follows where the heat-transfer rate is given by + pdv dφ ) mc dt Vdφ + pdv dφ The heat-transfer coefficient h c uses the correlation formula given by Woschni. 16 C p and C V are temperature-dependent parameters; their formulas are given in ref 16. The primary source is cylinder pressure-crank angle data. Using those primary data and the above formula, we can calculate peak (16) Heywood, J B. Internal Combustion Engine Fundamentals; McGraw- Hill: New York, 1988. (1) pv ) mrt (2) p dv dφ + Vdp dφ ) mrdt dφ dφ ) pc p dv R dφ + C V V dp R dφ + dq W dφ (3) (4) dq W dφ ) h c A(T - T W ) (5)
Combustion Characteristics at Various Injection Times Energy & Fuels, Vol. 20, No. 4, 2006 1501 Figure 5. timings. Excessive-air ratio of fuel blends versus fuel-injection pressure p max, mean gas temperature T, and maximum mean gas temperature T max. The flame-development duration is the angle interval from ignition start to the angle at which 10% accumulated heat release is reached; the rapid-combustion duration is the angle interval from 10% accumulated heat release to the angle of 90% accumulated heat release; the total-combustion duration is the angle interval from the beginning of heat release to the ending of heat release. The crank angle of the center of heat release curve is determined by the following formula φ e φs dφ φdφ φ c ) φ e φs dφ dφ in which φ s is the crank angle at the beginning of heat release and φ e is the crank angle at the end of heat release. Results and Discussions Figure 5 gives the excessive-air ratio versus fuel-injection timings. For a specific hydrogen fraction, the excessive-air ratio decreases with advancing fuel-injection timings. For a specific injection timing, the excessive-air ratio increases with an increase in hydrogen fraction. Because injection duration and injection pressure are kept constant, the volumetric amount of injected fuel per cycle remains the same. The advancement of fuel-injection timing reduces the amount of fresh air flowing into the cylinder and thus brings a decrease in the excessive-air ratio and richens the mixtures. The stoichiometric equivalence ratio of hydrogen is about one-fourth that of natural gas; the addition of hydrogen into natural gas increases the excessiveair ratio and makes the mixtures lean. Figure 6 gives the brake mean effective pressure (bmep) versus fuel-injection timings. The figure shows that bmep increases with advancing fuel-injection timings, reaches a maximum value at an injection timing of 190 CA BTDC, and maintains this high value with further advancement of the fuelinjection timings. A linear increase in bmep with the advancement of fuel-injection timings is presented when fuel-injection timings are less than 190 CA BTDC. The reason is that reducing the time interval between fuel injection and ignition decreases the time period for the fuel to become well-mixed, which causes the mixture in the cylinder to be nonuniform. Meanwhile, reducing the time interval between fuel injection and ignition decreases the penetration distance of the fuel jet because of an increased cylinder pressure due to piston compression. This leaves more injected fuel at the region near (6) Figure 6. Brake mean effective pressure versus fuel-injection timings. the nozzle tip, which leads to the formation of a rich stratified mixture in the local region of the combustion chamber and a lean mixture in the other region. Advancement of the fuelinjection timing supplies more time for fuel-air mixing, which improves the quality of the fuel-air mixing, shortens the combustion duration, and increases the cylinder pressure. Better fuel-air mixing is achieved at a fuel-injection timing of 190 CA BTDC, giving a high value of bmep. The study also shows that further advancement of fuelinjection timing (fuel-injection timing is less than 190 CA BTDC) has little influence on bmep, which can be explained by the following reasons. Enough time for fuel-air mixing is provided when the fuel-injection timing is less than 190 CA BTDC; thus, combustion will depend on the degree of mixture stratification and mixture concentration. Mixture stratification will weaken with further advancement of the fuel-injection timing at the moment of ignition starting (the time interval between the injection timing and the ignition timing is increased); this will decrease the burning velocity of the mixture, increasing the combustion duration and lowering the bmep. However, further advancement of the fuel-injection timing also decreases the amount of fresh air and excessive-air ratio, as demonstrated in Figure 5. This will increase the mixture burning velocity, shorten the combustion duration, and increase the bmep. The comprehensive effects of these two factors determine the behavior of bmep with further advancement of the fuelinjection timing. The experimental results show that these two opposite influences become offsetting and give little variation in bmep with further advancement of the fuel-injection timing. For a specific injection timing, an increase in hydrogen fraction decreases bmep when the hydrogen fraction is less than 10%, whereas bmep tends to increase when the hydrogen fraction is larger than 10%. The addition of hydrogen into natural gas in the direct-injection engine has two influencing aspects in combustion. In one aspect, because the intake fresh air is invariable at the fixed engine speed, the addition of hydrogen into natural gas will increase the excessive-air ratio of the mixture; the excessive-air ratio of the direct-injection engine is relatively high, and thus an increase in the excessive-air ratio will decrease the burning speed of the mixture from the view of mixture dilution. On the other hand, the addition of hydrogen into natural gas can increase the burning speed of the mixture from the view of hydrogen enrichment, which has been verified in the previous work in both combustion bomb and engine studies. As explained above, the influence gives the opposite trend in burning velocity (mixture dilution and hydrogen enrichment).
1502 Energy & Fuels, Vol. 20, No. 4, 2006 Huang et al. The comprehensive effect will determine the final result by hydrogen addition. A critical percentage of hydrogen will exist because of the combined effect, and the critical percentage of 10% hydrogen in volume fraction is experimentally obtained in the study. When the hydrogen fraction is less than 10%, the decrease in the burning velocity by mixture dilution is larger than the increase in burning velocity by hydrogen enrichment; this leads to a decrease in bmep. Vice versa, when the hydrogen fraction is larger than than 10%, the increase in burning velocity by hydrogen enrichment is larger than the decrease in burning velocity by mixture dilution; this increases the bmep. This reveals that the addition of hydrogen in natural gas can improve the combustion process when the hydrogen fraction is larger than a certain value under the condition of fixed injection duration. In addition, a similar trend of bmep with fuel-injection timing is presented for both natural gas direct-injection combustion and natural gas-hydrogen direct-injection combustion, indicating the same sensitivity of bmep with variation of fuelinjection timing. Figure 7 illustrates the combustion durations versus fuelinjection timings for natural gas direct-injection combustion and natural gas-hydrogen direct-injection combustion. Flamedevelopment duration decreases with advancing fuel-injection timing for both natural gas combustion and natural gashydrogen combustion, as shown in Figure 7a. This is due to the improvement of fuel-air mixing at early injection timing. For specific fuel-injection timing, natural gas-hydrogen directinjection combustion gives a longer flame-development duration than that of natural gas direct-injection combustion; which is due to the increase in the excessive-air ratio and dilution of the mixtures. In the case of natural gas-hydrogen direct-injection combustion, the flame-development duration decreases with an increase in hydrogen fraction; this is regarded as an improvement in ignitibility with an increase in hydrogen fraction. Furthermore, the difference in flame-development durations among different fuels is decreased with advancing fuel-injection timing, where weak mixture stratification leads to the decrease in flamedevelopment duration. Rapid-combustion duration (Figure 7b) and total-combustion duration (Figure 7c) decrease with advancing fuel-injection timing, and natural gas direct-injection combustion gives shorter combustion durations than those of natural gas-hydrogen directinjection combustion. When the hydrogen fraction is less than 10%, both rapid-combustion duration and total-combustion duration increase with an increase in hydrogen fraction. As explained above, an increase in the excessive-air ratio and dilution in mixture are responsible for this. Decreasing rapidcombustion duration and total-combustion duration is observed when the hydrogen fraction is larger than 10%, which is due to the remarkable increase in burning velocity at large hydrogen fractions. Similarly, the difference in rapid-combustion duration and total-combustion duration between natural gas directinjection combustion and natural gas-hydrogen direct-injection combustion decreases with advancing fuel-injection timing; even short-combustion durations are presented for natural gashydrogen combustion at a fuel-injection timing of 210 CA BTDC. Figure 8 gives the crank angle of the center of heat release curve (φ c ) versus fuel-injection timing. φ c reflects the heatrelease compactness that is usually used as a parameter for indicating combustion characteristics. The figure shows that φ c decreases with advancing fuel-injection timings and that more compactness in the heat-release process is presented at early fuel-injection timings. This can also be explained by the Figure 7. Combustion durations versus fuel-injection timings. influence from the variations in the excessive-air ratio and mixing quality. φ c increases with an increase in hydrogen fraction when the hydrogen fraction is less than 10% and decreases with any further increase in hydrogen fraction when the hydrogen fraction is larger than 10%. This is consistent to the behavior of combustion duration. Figure 9 shows the maximum cylinder pressure and maximum mean gas temperature versus fuel-injection timings for natural gas direct-injection combustion and natural gas-hydrogen direct-injection combustion. Figure 10 gives NO x, CO, and CO 2 concentrations versus fuelinjection timings for natural gas direct-injection combustion and natural gas-hydrogen direct-injection combustion. NO x increases with advancing fuel-injection timing; this is due to mixing improvement and better combustion, which increases the combustion temperature and decreases the excessive-air ratio. These two factors lead to the increase in NO x concentration.
Combustion Characteristics at Various Injection Times Energy & Fuels, Vol. 20, No. 4, 2006 1503 Figure 8. Crank angle of the center of heat release curve versus fuelinjection timings. Figure 10. Exhaust-gas concentrations versus fuel-injection timings. Figure 9. P max and T max versus fuel-injection timings. For a specific fuel-injection timing, the NO x concentration of natural gas-hydrogen direct-injection combustion gives a lower value than that of natural gas combustion, which is due to the dilution of the mixture when hydrogen is added in natural gas at a fixed injection duration. The exhaust CO concentration experiences small variations under various fuel-injection timings. This is reasonable, because the excessive-air ratios in this study are greater than 1 and the CO concentration remains very low at lean mixture combustion. The exhaust CO 2 increases with advancing fuel-injection timing, which is a result of the improvement in fuel-air mixing and decrease in the excessiveair ratio; the latter reduces the amount of cylinder charge and increases combustion gas concentrations. For a specific fuelinjection timing, the addition of hydrogen in natural gas can reduce CO 2 concentration, which is due to the decrease in the C:H ratio when hydrogen is added in natural gas. The study indicates that the addition of hydrogen in natural gas can reduce CO 2 concentration. Conclusions Combustion characteristics under various injection timings of a direct-injection engine fueled with natural gas-hydrogen blends at fixed injection duration and fixed ignition timing were investigated, and the main results are summarized as follows: 1. Brake mean effective pressure increases with advancing fuel-injection timings. Brake mean effective pressure reaches a maximum value at an injection timing of 190 CA BTDC and maintains this high value with further advancement of fuelinjection timings. For a specific injection timing, an increase in the hydrogen fraction leads to a decrease in the brake mean effective pressure when the hydrogen fraction is less than 10%,
1504 Energy & Fuels, Vol. 20, No. 4, 2006 Huang et al. whereas the brake mean effective pressure tends to increase when the hydrogen fraction is larger than 10%. 2. Combustion durations decrease with advancing fuelinjection timing. When the hydrogen fraction is less than 10%, combustion durations increase with an increase in hydrogen fractions. Combustion durations decrease with an increase in hydrogen fractions when the hydrogen fraction is larger than 10%. 3. NO x and CO 2 concentrations increase with advancing fuelinjection timing, and the CO concentration experiences small variations under various fuel-injection timings. The addition of hydrogen in natural gas can reduce the CO 2 concentration. Acknowledgment. The study was supported by the National Natural Science Foundation of China (50422261, 50521604), the National Basic Research Project (2003CB214501), and the State Key Laboratory Award Fund from the Natural Science Foundation of China (50323001). We acknowledge the students of Xi an Jiaotong University for their help with the experiment and preparation of the paper. We also express our thanks to our colleagues at Xi an Jiaotong University for their helpful comments and advice during paper preparation. A ) wall area (m 2 ) ATDC ) after top-dead-center Nomenclature BTDC ) before top-dead-center C p ) constant pressure specific heat (kj kg -1 K -1 ) C V ) constant volume specific heat (kj kg -1 K -1 ) ( /dφ) ) Heat-release rate with crank angle (J/ CA) (dq w /dφ) ) Heat-transfer rate with crank angle (J/ CA) h c ) heat-transfer coefficient (J m -2 s -1 K -1 ) H u ) heating value (MJ/kg) m ) mass of cylinder gases (kg) p ) cylinder gas pressure (MPa) p max ) maximum cylinder gas pressure (MPa) R ) gas constant (J kg -1 K -1 ) T ) mean gas temperature (K) T max ) maximum mean gas temperature (K) T w ) wall temperature (K) TDC ) top-dead-center V ) cylinder volume (m 3 ) φ ) crank angle (degrees) φ c ) crank angle of the center of heat release curve ( CA ATDC) φ e ) crank angle of heat-release ending ( CA ADTC) φ s ) crank angle of heat-release beginning ( CA BTDC) φ ign ) ignition advance angle ( CA BTDC) φ inj ) injection advance angle ( CA BTDC) EF060032T