Combustion Characteristics and Heat Release Analysis of a Spark

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Energy & Fuels 2007, 21, 2594-2599

Combustion Characteristics and Heat Release Analysis of a Spark-Ignited Engine Fueled with Natural Gas-Hydrogen Blends Zuohua Huang,* Bing Liu, Ke Zeng, Yinyu Huang, Deming Jiang, Xibin Wang, and Haiyan Miao State Key Laboratory of Multiphase Flow in Power Engineering, Xi’an Jiaotong UniVersity, Xi’an, People’s Republic of China ReceiVed March 28, 2007. ReVised Manuscript ReceiVed June 29, 2007

Combustion characteristics and heat release analysis of a spark-ignited engine fueled with natural gashydrogen blends were investigated. For the same excess air ratio and at the optimum ignition timing setting, the natural gas combustion gives higher values of peak cylinder pressure and peak heat release rate in the case of the rich mixture combustion. In the case of the stoichiometric mixture combustion and the lean mixture combustion, the natural gas-hydrogen blend with a 10% hydrogen fraction gives the highest value of the peak cylinder pressure and the peak heat release rate. The initial combustion duration and the total combustion duration decrease with the increase of hydrogen fraction in the natural gas-hydrogen blend. The addition of hydrogen into natural gas decreases the ignition delay. Although the optimum ignition timing is retarded with the increase of the hydrogen fraction in the natural gas-hydrogen blends, the heat release process is not postponed.

Introduction With increasing concern about energy shortages and environmental protection, research on improving engine fuel economy and reducing exhaust emissions has become a major aspect in combustion and engine development. Because of the limited reserves of crude oil, the development of alternative fuel engines has attracted more and more concern in the engine research community. Alternative fuels usually belong to clean fuels as compared to diesel fuel and gasoline fuel in the engine combustion process. The introduction of these alternative fuels is beneficial to slow down the fuel shortage and reduce the engine exhaust emissions. Natural gas is regarded as one of the favorable fuels for engines, and the natural gas fueled engine has been used in both the spark-ignited engine and the compression-ignited engine. However, due to the slow-burning velocity of natural gas and the poor lean-burn capability, the natural gas spark-ignited engine still has its disadvantages, such as the relatively large cycle-by-cycle variations in the lean mixture combustion and a poor lean-burn capability, and these restrict its operation range and increase its fuel consumption.1-2 Because of these restrictions, the natural gas engine is usually operated at conditions of a stoichiometric equivalence ratio with a relatively low thermal efficiency. Traditionally, to improve the lean-burn capability and flame-burning velocity of the natural gas engine under lean-burn conditions, an increase of flow intensity in the cylinder is needed, and this measure would increase the heat loss to the cylinder wall and increase the * Corresponding author. E-mail: [email protected]; fax: 008629-82668789. (1) Rousseau, S.; Lemoult, B.; Tazerout, M. Combustion characteristics of natural gas in a lean burn spark-ignition engine. Proc. Inst. Mech. Eng., Part D 1999, 213, 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, 1999.

combustion temperature and NOx emissions.3 One effective method to increase the burning velocity of a mixture is to mix the natural gas with fuel that possesses a fast-burning velocity. Hydrogen is regarded as the best fuel additive for natural gas due to its very fast-burning velocity, and this combination could improve the lean-burn characteristics and decrease the engine emissions.4 Blarigan and Keller investigated the port-injection engine fueled with natural gas-hydrogen mixtures.5 Wong and Karm6 and Huang et al.7 studied the engine performance fueled by various hydrogen fractions in natural gas-hydrogen blends. Bauer and Forest conducted an experimental study on natural gas-hydrogen combustion in a CFR engine.8 Furthermore, studies on the lean combustion capability of natural gashydrogen combustion and natural gas-hydrogen combustion with turbo-charging and/or exhaust gas recirculation were conducted,9-11 and these studies showed that the exhaust HC, (3) Das, A.; Watson, H. C. Development of a natural gas spark ignition engine for optimum performance. Proc. Inst. Mech. Eng., Part D 1997, 211, 361-378. (4) Akansu, S. O.; Dulger, A.; Kahraman, N. Internal combustion engines fueled by natural gas-hydrogen mixtures. Int. J. Hydrogen Energy 2004, 29, 1527-1539. (5) Blarigan, P. V.; Keller, J. O. A hydrogen fueled internal combustion engine designed for single speed/power operation. Int. J. Hydrogen Energy 2002, 23, 603-609. (6) 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, 12171224. (7) Huang, Z. H.; Liu, B.; Zeng, K.; Huang, Y. Y.; Jiang, D. M.; Wang, X. B.; Miao, H. Y. Experimental study on engine performance and emissions for an engine fueled with natural gas-hydrogen mixtures. Energy Fuels 2006, 20, 2131-2136. (8) Bauer, C. G.; Forest, T. W. Effect of hydrogen addition on the performance of methane-fueled vehicles. Part I: Effect on SI engine performance. Int. J. Hydrogen Energy 2001, 26, 55-70. (9) Sierens, R.; Rosseel, E. Variable composition hydrogen-natural gas mixtures for increased engine efficiency and decreased emissions. Trans ASME: J. Eng. Gas Turbines Power 2000, 122 135-140.

10.1021/ef0701586 CCC: $37.00 © 2007 American Chemical Society Published on Web 08/15/2007

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Table 1. Fuel Properties of Natural Gas and Hydrogen types of fuel (kg/m3)

density at 1 atm and 300 K stoichiometric A/F ratio (vol %) laminar-burning velocity (m/s) quenching distance (mm) conductivity at 300 K (mW/m2 K) lower heating value (MJ/kg) lower heating value (MJ/m3) cetane number C/H ratio

Table 2. Compositions of Natural Gas

natural gas

hydrogen

Items

CH4

C2H6

C3H8

N2

CO2

other

0.754 9.396 0.38 2.03 34 43.726 32.97 127 0.252

0.082 2.387 2.9 0.64 182 119.7 9.82

volumetric fraction (%)

96.160

1.096

0.136

0.001

2.540

0.067

0

CO, and CO2 concentrations could be decreased when an engine operated on natural gas-hydrogen. However, NOx may increase in the natural gas-hydrogen combustion at rich mixture conditions due to an increase in the flame propagation speed and gas temperature. The NOx concentration can be greatly decreased through lean combustion and retarding the ignition timing. Up to now, most previous work has concentrated on engine performance and emissions fueled with homogeneous natural gas-hydrogen mixtures at stoichiometric equivalence ratios, and few literature has reported engine performance and emissions at various equivalence ratios when using natural gashydrogen blends. Shudo et al. investigated the combustion and emissions of an engine with port-injected hydrogen and incylinder injection natural gas;12 however, this type of engine needs two separate fueling systems and makes the system complicated. This paper investigates the performance and emissions of a spark-ignited engine fueled with various fractions of natural gas-hydrogen mixtures and operated at various equivalence ratios. The study will clarify the behavior of an engine fueled with various fractions of natural gas-hydrogen mixtures and at various equivalence ratios. The study will enhance the understanding of a natural gas-hydrogen fueled engine and provide practical guidance to engine application. Experimental Procedures A three-cylinder automotive CNG spark-ignited engine was used, and the specifications of the engine are as follows: cylinder bore of 68.5 mm, stroke of 72 mm, displacement of 796 mL, compression ratio of 9.4, and rated speed and power of 5500 rpm and 26.5 kW, respectively. Natural gas-hydrogen blends with different hydrogen fractions were prepared in advance in the CNG tank. In this study, four fuel blends were prepared: the pure natural gas, the blend with 90% natural gas and 10% hydrogen by volume, the blend with 80% natural gas and 20% hydrogen, and the blend with 74% natural gas and 26% hydrogen. Fuel was supplied to the engine intake port through a gas mixer, and the amount of fuel and excess air ratios were controlled by an ECU unit and were regulated by the stepmotor in the mixer. Experiments were conducted at WOT, the optimum ignition timing or MBT, and an engine speed of 2000 rpm. Table 1 gives fuel properties of natural gas and hydrogen. Hydrogen with a purity of 99.995% was used, while the natural gas composition is listed in Table 2. It can be seen that the laminarburning velocity of hydrogen is 5 times that of natural gas and that the quenching distance of hydrogen is one-third that of natural gas, while the latter is beneficial to reduce the unburned hydrocarbons (10) Larsen, J. F.; Wallace, J. S. Comparison of emissions and efficiency of a turbocharged lean-burn natural gas and hythane-fueled engine. Trans ASME: J. Eng. Gas Turbines Power 1997, 218, 218-226. (11) 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. Proc. Inst. Mech. Eng., Part D 2001, 215, 405-418. (12) Shudo, T.; Shimamura, K.; Nakajima, Y. Combustion and emissions in a methane DI stratified charge engine with hydrogen pre-mixing. JSAE ReV. 2000, 21, 3-7.

near the wall and from the top-land crevice. Meanwhile, hydrogen has a higher conductivity than that of natural gas, and this may increase the heat loss to the coolant in the case of natural gashydrogen combustion as compared to that of natural gas combustion. The mass lower heating value of hydrogen is larger than that of natural gas, but the volumetric lower heating value of hydrogen is smaller than that of natural gas. Instrumentation and Method of Calculation. The cylinder pressure was recorded by a piezoelectric transducer with a resolution of 10 Pa, and the dynamic TDC was determined under the motoring operation. The crank angle signal was obtained from the angle generating device mounted on the main shaft. The signal of cylinder pressure was acquired for every 0.5 deg CA, the acquisition process covered 254 completed cycles, and the averaged value of these 254 cycles was outputted as the pressure data for calculation of the combustion parameters. Brake mean effective pressure (bmep) was calculated from the engine power, speed, and configuration specifications. A thermodynamic model was used to calculate the thermodynamic parameters in this paper, and the model neglects the leakage through the piston rings;13 thus, the energy conservation in cylinder is written as follows: dQB dQW d(mu) dCV dV dV dT ) + p ) mCV + mT +p (1) dφ dφ dφ dφ dφ dφ dφ The gas-state equation is pV ) mRT

(2)

The variation of the gas-state equation with the crank angle is given by dV dp dT p + V ) mR dφ dφ dφ

(3)

The heat release rate dQB/dφ can be derived from eqs 1 and 3 as follows: dQB Cp dV CVV dp dCV dQW )p + + mT + dφ R dφ R dφ dφ dφ

(4)

where the heat transfer rate is given by dQW ) hcA(T - TW) dφ

(5)

The heat transfer coefficient hc uses the correlation formula given by Woschni in ref 14. Cp and CV are temperature-dependent parameters, and their formulas are given in ref 14. Figure 1 gives the volumetric heating value of the natural gashydrogen-air mixtures versus the hydrogen fractions at the stoichiometric equivalence ratio. It can be seen that the volumetric heating value of the natural gas-hydrogen mixture decreases with the increase of hydrogen fraction in the blend, and this would be due to the low volumetric heating value of the hydrogen-air mixture as compared to that of the natural gas-air mixture at the stoichiometric equivalence ratio. Thus, for a given fuel injection duration, the amount of heat release decreases with the increase of hydrogen fraction in the blends. To maintain the same equivalence (13) Huang, Z. H.; Shiga, S.; Ueda, T.; Jingu, N.; Nakamura, H.; Ishima, T.; Obokata, T.; Tsue, M.; Kono, M. A basic behavior of CNG DI combustion in a spark-ignited rapid compression Machine. JSME Int. J., Ser. B 2002, 45, 891-900. (14) Heywood, J. B. Internal Combustion Engine Fundamentals; McGrawHill Book Company: New York, 1988.

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Figure 1. Lower heating value of fuel blends vs hydrogen fractions.

Figure 2. C/H ratio of fuel blends vs hydrogen fractions.

Figure 3. Optimum ignition timing vs excess air ratio.

ratio, more fuel needs to be injected in the case of the natural gashydrogen mixture combustion. The volumetric heating value of the natural gas-hydrogen mixture decreases with the increase of the hydrogen fraction, and this suggests that more fuel volumetrically needs to be supplied to the engine to maintain the same heat release per cycle with the increase of the hydrogen fraction. Figure 2 shows the ratio of C/H versus the hydrogen fraction. The C/H ratio decreases with the increase of the hydrogen fraction in the natural gas-hydrogen mixture. For example, the C/H ratio decreases by 11% in the case of 20% hydrogen addition. Lowering the C/H ratio is beneficial to the reduction of carbon-related emissions, such as CO, CO2, and unburned hydrocarbons.

Results and Discussion Figure 3 gives the optimum ignition timings versus the excess air ratio for various hydrogen fractions in fuel blends. The optimum ignition timing is the ignition timing to obtain the MBT. For all mixtures, the optimum ignition timing is advanced with the increase of excess air ratio. The optimum ignition timing will be retarded with the increase of hydrogen addition in the natural gas-hydrogen blend. A lean mixture with a large excess air ratio decreases the burning velocity, and this needs an advancement of ignition timing to prevent the loss of power due to long combustion duration. The addition of hydrogen in natural gas can shorten the ignition delay and increase the

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burning velocity of the mixture. The shortening in the ignition delay retards the ignition timing and avoids the large rising of the cylinder pressure at the late stage of the compression stroke. In the case of natural gas-hydrogen combustion, appropriate retarding in the ignition timing does not postpone the combustion process since the burning velocity of the mixture increases by hydrogen addition. The MBT setting is presented at a stoichiometric equivalence ratio (λ ) 1.0) or an equivalence ratio slightly less than the stoichiometric equivalence ratio (λ ) 0.9) for both the natural gas combustion and the natural gashydrogen combustion. The figure also shows that the lean-burn limit can be extended with the increase of the hydrogen fraction in the blend, and this is due to the improvement of the mixture ignitibility and combustion when hydrogen is added. Figure 4 gives the cylinder pressure of blends with different hydrogen fractions and at four excess air ratios. In the case of the rich mixture combustion (λ ) 0.9), the natural gas combustion gives an early rising in cylinder pressure and a higher value of peak pressure. With the increase of the hydrogen fraction in natural gas-hydrogen blends, the rising point in the cylinder pressure is retarded, and the crank angle at the peak pressure is postponed, leading to the decrease of the peak cylinder pressure with the increase of the hydrogen fraction in natural gas-hydrogen blends. Although the flame propagation speed is increased with the increase of the hydrogen fraction in the natural gas-hydrogen blends, the retarding in the optimum ignition timing with the increase of the hydrogen fraction still postpones the rising of the cylinder pressure. In addition, the decrease in the heating value of the blend with the increase of the hydrogen fraction decreases the peak value of the cylinder pressure. In the case of the stoichiometric mixture combustion (λ ) 1.0) and the lean mixture combustion (λ ) 1.2 and 1.4), the blend with the 10% hydrogen fraction gives the earliest cylinder pressure rising and the highest value of the peak cylinder pressure. When further increasing the hydrogen fraction in natural gas-hydrogen blends, the rising point in the cylinder pressure is postponed, accompanied by decreasing the peak cylinder pressure and retarding the crank angle of the peak cylinder pressure. The figure also shows that the differences among the blends at different hydrogen fractions become small when increasing the excess air ratio and/or using a leaner mixture combustion. The study reveals that, at the same excess air ratio and optimum ignition timing setting, the influence from hydrogen addition into natural gas can be reflected at the stoichiometric mixture combustion and the lean mixture combustion. Two factors are considered to influence the cylinder pressure: one is the increase in flame propagation speed or combustion speed with the increase of the hydrogen fraction in the blends, and this will cause a rapid rising in the cylinder pressure and bring a higher value of the peak cylinder pressure; another is the decrease in the heating value of the fuel blends with the increase of the hydrogen fraction in natural gashydrogen blends, and this will decrease the volumetric heat release rate and the cylinder pressure rising, leading to the lower value of the peak cylinder pressure. When the hydrogen fraction is less than 10%, the effect from the increase in the flame propagation speed is larger than that from the decrease in the heat value of the blends, and this leads to an increase of the cylinder pressure with the increase of the hydrogen fraction. When the hydrogen fraction is larger than 10%, the influence from the flame propagation speed promotion is lower than the influence from the decrease in the heat value of the blends, and this leads to the decrease of the cylinder pressure with the

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Figure 4. Cylinder pressure at different hydrogen fractions.

Figure 5. Heat release rate at different hydrogen fractions.

increase of the hydrogen fraction. Since the difference in the heat value of the fuel-air mixture becomes smaller in the case of the lean mixture, the cylinder pressure at the lean mixture combustion gives a small difference among various natural gashydrogen blends. Figure 5 shows the heat release rate of the blends with different hydrogen fractions at four excess air ratios. Similar to those of the cylinder pressure, in the case of the rich mixture combustion (λ ) 0.9) and the optimum ignition timing, the heat release rate decreases with the increase of the hydrogen fraction

in the natural gas-hydrogen blends. In the case of the stoichiometric mixture combustion (λ ) 1.0) and the lean mixture combustion (λ ) 1.2 1.4), the fastest heat release rate and the highest value of the maximum heat release rate are presented for the blend with the 10% hydrogen addition, while further increasing the hydrogen fraction will decrease the maximum heat release rate. As interpreted previously, the combined influence from the burning velocity promotion and the decrease in the heating value of the mixture with the increase of the hydrogen fraction is responsible for this behavior.

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Figure 7. Rapid combustion duration.

Figure 6. Initial combustion duration.

Although optimum ignition timing is retarded with the increase of the hydrogen fraction, little postponement in the heat release is observed, and this is due to the decrease in the ignition delay and the rapid combustion in the case of the natural gashydrogen combustion. In this paper, the initial combustion duration is defined as the interval of the crank angle from the ignition start to that of 10% accumulated heat release, the rapid combustion duration is defined as the interval of crank angle from 10% accumulated heat release to that of 90% accumulated heat release, and the sum of the initial combustion duration and the rapid combustion duration is defined as the total combustion duration. Figure 6 shows the initial combustion duration of the natural gas-hydrogen blends at various excess air ratios and hydrogen fractions. For a specified fuel blend, the initial combustion duration gives the shortest duration at λ ) 1.1, and both increasing and decreasing the excess air ratio will increase the initial combustion duration. For a given excess air ratio, the initial combustion duration shows a linear decreasing trend with the increase of hydrogen fraction in the natural gashydrogen blends. The availability of hydrogen in the mixture improves the mixture ignitability and promotes flame early development. Thus, hydrogen addition is beneficial to the mixture ignitibility and the flame early development, in other words, a shorter ignition delay can be realized with the addition of hydrogen. Figure 7 gives the rapid combustion duration versus hydrogen fraction, and the total combustion duration versus hydrogen fraction is illustrated in Figure 8. For a given excess air ratio, the rapid combustion duration shows a slight increase with the increase of the hydrogen fraction when the hydrogen fraction is less than 20%, and the rapid combustion duration decreases with the increase of the hydrogen fraction when the hydrogen fraction is larger than 20%. Two factors are considered to bring this phenomenon: one is the addition of hydrogen, which improves the burning speed of mixture; another is the lowering in the volumetric heating value of the mixture by hydrogen addition at the same excess air ratio, and this may reduce the

Figure 8. Total combustion duration.

burning velocity of the mixture. This is consistent with the results obtained in a constant volume vessel study where the burning velocity gave a slight increase with the increase of the hydrogen fraction at a hydrogen fraction less than 20%, while a rapid increase in the burning velocity was presented when the hydrogen fraction was over 20%.15 The total combustion duration decreases slightly with the increase of the hydrogen fraction at hydrogen fractions less than 20%, while a rapid decreasing is presented at hydrogen fractions over 20%. In the case of the leaner mixture operation (λ ) 1.6), the flame propagation speed is strongly dependent on the burning speed due to hydrogen addition. As the mixture becomes leaner, the influence from the decreased heating value by hydrogen addition becomes small. The lean mixture of the natural gas has a slow flame propagation speed; thus, the addition of hydrogen will become a dominant factor. Figure 9 shows the maximum mean gas temperature versus excess air ratio and the hydrogen fraction. For a specified blend, the maximum mean gas temperature decreases with the increase of excess air ratio due to mixture dilution. For a given excess air ratio, the maximum mean gas temperature increases slightly with the increase of the hydrogen fraction at hydrogen fractions less than 20%, and a slight dropping in the maximum mean gas temperature is presented when the hydrogen fraction is over 20%. The comprehensive influence from the improvement in burning velocity and the lowering in the fuel heating value is responsible for this phenomenon. The improvement in burning velocity tends to increase the gas temperature while the lowering in the fuel heating value tends to decrease the gas temperature. In the case of an excess air ratio of 1.6, the maximum mean gas temperature increases with the increase of the hydrogen fraction. (15) Huang, Z. H.; Zhang, Y.; Zeng, K.; Liu, B.; Wang, Q.; Jiang, D. M. Measurements of laminar burning velocities for natural gas-hydrogenair mixtures. Combust. Flame 2006, 146, 302-311.

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fraction gives the highest values of peak cylinder pressure and maximum heat release rate. (2) For the same excess air ratio and at optimum ignition timing, the initial combustion duration and the total combustion duration decrease with the increase of the hydrogen fraction the in natural gas-hydrogen blends. (3) The addition of hydrogen into natural gas decreases the ignition delay. Although the optimum ignition timing is retarded with the increase of the hydrogen fraction in the natural gashydrogen blend, the heat release process is not postponed. Acknowledgment. This study was supported by the NSFC Fund (50636040, 50521604, and 50323001) and the National Basic Research Project (2003CB214501).

Nomenclature A ) wall area ATDC ) after top-dead-center BTDC ) before top-dead-center Cp ) constant pressure specific heat (kJ/kg K) Cv ) constant volume specific heat (kJ/kg K) dQB/dφ ) heat release rate with crank angle (J/crank angle (CA)) dQw/dφ ) heat transfer rate with crank angle (J/CA) hc ) heat transfer coefficient (J/m2 s K) Hu ) lower heating value (MJ/kg) m ) mass of cylinder gases (kg) MBT ) minimum spark advance for best torque (CA deg BTDC) p ) cylinder gas pressure (MPa) R ) gas constant (J/kg K) T ) mean gas temperature (K) Tmax ) maximum mean gas temperature (K) Tw ) wall temperature (K) TDC ) top-dead-center V ) cylinder volume (m3) WOT ) wide open throttle θfd ) initial combustion duration (CA deg) θrd ) rapid combustion duration (CA deg) λ ) excess air ratio (m2)

Figure 9. Maximum mean cylinder gas temperature.

Conclusion Combustion characteristics and heat release analysis of a spark-ignited engine fueled with natural gas-hydrogen blends were investigated, and the main results are summarized as follows: (1) for the same excess air ratio and at optimum ignition timing, the natural gas combustion gives the high values of peak cylinder pressure and maximum heat release rate in the case of the rich mixture combustion. In the case of the stoichiometric mixture combustion and the lean mixture combustion, the natural gas-hydrogen blend with a 10% hydrogen

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