Seoul FISIT World utomotive Congress June 1-15,, Seoul, Korea F85 nalysis of Direct Injection Spark Ignition Combustion in Hydrogen Lean Mixture Toshio Shudo 1) * Koichiro Tsuga ) Yasuo Nakajima 1) 1) Musashi Institute of Technology, Tamazutsumi 1-8-1, Setagaya-ward, Tokyo, Japan ) Musashi Institute of Technology, Graduate School, Tamazutsumi 1-8-1, Setagaya-ward, Tokyo, Japan Characteristics of methane direct injection spark ignition stratified combustion in lean hydrogen mixture were analyzed both in a single cylinder engine and in a constant volume combustion vessel. Combustion pressure and instantaneous combustion chamber wall temperature during combustion process were measured and used in analysis of combustion and cooling loss. Results in this research show the premixed hydrogen increases cooling loss to combustion chamber wall while achieving combustion promotion, and the combustion system is effective especially in lean mixture conditions. nalysis of flame propagation was also done with schlieren photography in the constant volume combustion vessel. Keywords: Stratified Charge, Hydrogen, Methane, Lean Combustion, Cooling Loss INTRODUCTION lthough direct injection natural gas engines are expected to have higher thermal efficiency with lean combustion, there remain some problems such as low combustion velocity and relatively high unburned hydrocarbon exhaust emission at light load lean operation. Direct injection stratified charge combustion with twostage fuel injection system has been reported to promote combustion and to expand range of knock free operation in the use of hydrocarbon fuels [1-5]. On the other hand, hydrogen has an extremely higher burning velocity and wider flammable limits compared with hydrocarbon fuels []. Hydrogen addition to methane has been reported to be effective to promote combustion at lean operation [7-8]. In this research, characteristics of methane direct injection spark ignition stratified combustion in lean hydrogen mixture were analyzed in a single cylinder engine and a constant volume combustion vessel. Combustion pressure and instantaneous combustion chamber wall temperature during combustion process were measured and used in analysis of combustion characteristics and cooling loss to combustion chamber wall. Results in this research show that the combustion system can achieve lower hydrocarbon exhaust emission and higher thermal efficiency compared with methane direct injection spark ignition combustion, however the premixed hydrogen increases cooling loss to combustion chamber wall while achieving combustion promotion. The combustion system of methane direct injection stratified combustion in hydrogen lean mixture is effective especially in lean conditions. EXPERIMENT SINGLE CYLINDER ENGINE Tested engine was a four-stroke cycle single cylinder spark ignition engine with a bore stroke of 85 88mm and a compression ratio of 13. Hydrogen was continuously supplied into the intake manifold, and methane was directly injected into the cylinder with a gas injector at an injection timing of 35 C TDC with an injection pressure of 1MPa. The fuel injection direction was toward the spark plug as shown in Fig.1. For each experimental condition, cycles of in-cylinder pressure data were measured with piezoelectric type pressure transducer (VL GM1D) installed in the cylinder head and used to analyze indicated Fuel injector - φ85 Intake valve Spark plug Pressure transducer (VL GM1D) Exhaust valve Fuel jet Fig.1 Combustion chamber of tested single cylinder engine Spark plug 1 Electrom agnetic fuelinjector O bservation window s 8 Pressure transducer (Kistler 71) Thin film therm ocouple (chrom el-constantan, M edtherm TC S-13E) Fig. Constant volume combustion vessel in this research - * Corresponding author. e-mail: shudo@herc.musashi-tech.ac.jp 1
THC ppm 15 1 5 5 Engine speed : 1rpm Ignition timing : 15 C BTDC Local heat flux J/(m C) 1 8 Volumetric efficiency : 5% Engine speed : 15rpm λ=1. hydrogen Tig=18 BTDC Tig=9B Tig=TDC(MBT) Tig=9 Tig=18 NOx ppm 3 1-9 -5 5 9 Crank angle C TDC (a) Hydrogen combustion ηi % 3 5 15 1 5 only H premixed + H/CH = % H/CH = % 1 3 5 Overall excess air ratio Fig.3 Influence of overall excess air ratio on thermal efficiency and exhaust emissions thermal efficiency. NOx and THC exhaust emissions were measured with CLD and FID gas analyzer. CONSTNT VOLUME COMBUSTION VESSEL constant volume combustion vessel used in this research is shown in Fig.. The vessel had transparent quartz observation windows on both sides. Methane was directly injected into lean pre-mixture of hydrogen and air, and was ignited with a spark plug. The fuel injection direction was toward the spark plug. Methane was injected with an electromagnetic fuel injector installed with a nozzle (a hole of φ1mm). Fuel injection pressure was at 3MPa. Combustion pressure was measured with a piezoelectric type pressure transducer (Kistler 71) and used in analysis of heat release rate, heat release, and combustion period. Instantaneous temperature data of combustion chamber wall surface were measured with a thin film type thermo-couple (Medtherm TCS-13E,chromel-constantan type) and used in analysis of cooling loss characteristics. Observation of flame propagation was done with the Schlieren method using a mercury lump (Ushio USH-5D). Images were recorded with a memory type high speed video camera (Photoron FSTCM ultima). RESULTS ND DISCUSSION RESULTS OF ENGINE EXPERIMENTS Fig.3 is a result of engine experiment showing influence of hydrogen premixing ratio on combustion and exhaust emissions. Engine speed was set at 1rpm, and Local heat flux J/(m C) 1 8 Volumetric efficiency : 5% Engine speed : 15rpm λ=1. Methane Tig=5 BTDC Tig=3B Tig=7B(MBT) Tig=18B Tig=9B -9-5 5 9 Crank angle C TDC (b) Methane combustion Fig. Instantaneous heat flux of hydrogen combustion and methane combustion [9] (-stroke cycle spark ignition engine with cylinders. Bore stroke: 85 88mm. Compression ratio: 13.. Engine speed: 15rpm. Volumetric efficiency: 5%. Premixed stoichiometric combustion.) ignition timing was at 15 C TDC. The overall excess air ratio was calculated with both premixed hydrogen and directly injected methane. The hydrogen premixing improve thermal efficiency and reduces THC largely over the wide range of excess air ratio. lthough NOx increases especially at condition between stoichiometric to excess air ratio of., the hydrogen premixing enables operation at very lean conditions with NOx and THC emission. However, hydrogen combustion has higher cooling loss to combustion chamber wall than hydrocarbon [9]. Combustion velocity of hydrogen is around 7 times higher than hydrocarbons, and quenching distance is around 1/ of hydrocarbons. The high flame propagation velocity increases forced convection between flame and chamber wall. Short quenching distance decreases width of temperature boundary layer on chamber wall. Fig. indicates combustion chamber wall instantaneous heat flux of hydrogen combustion and methane combustion in a premixed combustion engine. Compared with methane combustion, hydrogen combustion has higher amount of heat flux during combustion period. It means that hydrogen combustion has higher amount of cooling loss. From these results, the premixed hydrogen is supposed to change cooling loss characteristics in the methane direct
Pressure MPa Rate of heat release kj/s 1..8.... 5 3 1 H prem ixed % 3% % 1% O verallexcess air ratio :1.5 H prem ixed 1% % 3% % -1 1 3 5 7 Time after ignition ms Fig.5 Influence of hydrogen addition on methane DI combustion Local heat flux W/m 1 8 H prem ixed % % O verallexcess air ratio :1.5-1 3 5 7 Time after ignition ms Fig. Influence of hydrogen addition on local heat flux injection stratified charge combustion. This research analyzed combustion promotion and cooling loss of methane direct injection stratified charge combustion with hydrogen premixing using the constant volume combustion vessel. INFLUENCE OF HYDROGEN PRENIXING RTIO Fig.5 shows influence of hydrogen premixing ratio on combustion characteristics. Experiments were done with the constant volume combustion vessel shown in Fig.. Overall excess air ratio that was calculated with both premixed hydrogen and directly injected methane was set at 1.5. Hydrogen premixing ratio in the figure shows proportion of premixed hydrogen to totally supplied fuel in low heating value. With increase in hydrogen premixing ratio, combustion was promoted, and combustion period is shortened. The increase in hydrogen premixing ratio causes decrease in total fuel heat. But the difference is.% at the most and its influence is negligible. Fig. shows instantaneous heat flux to combustion chamber that was calculated from the combustion chamber wall surface instantaneous temperature. The instantaneous heat flux was calculated with an assumption of onedimensional thermal conduction in the combustion chamber wall. With increase in hydrogen premixing ratio, the instantaneous local heat flux increases showing the increase in cooling loss to combustion chamber. These results were analyzed using following indexes; kj/s dq /dtmax θ1-9 ms Q/Qfuel M ax.localheat flux W/m 5 λ=1. 3 1 λ=1.5 1 8.8.7..5. 1 8 1 3 5 Prem ixed/totalfuellh V ratio % Fig.7 Influence of premixed hydrogen amount on combustion 1 maximum value of heat release rate: dq/dt max 1-9% combustion period: θ 1-9 3 maximum value of instantaneous heat flux: q max ratio of heat release to supplied fuel heat: Q /Q fuel Degree of combustion promotion was evaluated with maximum heat release rate dq/dt max and 1-9% combustion period θ 1-9. Tendency of cooling loss was evaluated with maximum value of instantaneous local heat flux q max and Q /Q fuel. The index Q /Q fuel was calculated with total heat release Q and low heating value of supplied fuel Q fuel. The Q calculated with pressure data is affected by cooling loss, and described using actual heat release Q B and cooling heat Q C as follows. Q = Q B - Q C Here, the actual heat release Q B is affected by combustion efficiency and is described as follows. Q B =η u Q fuel nd the Q /Q fuel can be described, Q /Q fuel =(Q B -Q C )/ Q fuel =(Q B -Q C ) η u / Q B =η u (1-φ w ) Where, φ w = Q C /Q B. Therefore, the Q/Q fuel corresponds to a function of combustion efficiencyη u and cooling loss ratioφ w which is defined above with actual heat release Q B and cooling heat 3
Pressure MPa 1..8.. λ=1. λ=1. λ=. λ=. Pressure MPa 1..8.. λ=1. λ=1... dq /dt kj/s. 5 3 1 λ=1. λ=1. λ=. λ=. dq /dt kj /s. 5 3 1 λ=1. λ=1. 1 3 5 7 Tim e after ignition m s 1 3 5 7 Tim e after ignition m s (a) H premixed % + methane DI (b) Methane DI only Fig.8 Influence of excess air ratio on combustion Localheat flux W/m 1 λ=1. 8 λ=. - 1 3 5 7 Tim e after ignition ms (a) H premixed % + methane DI (b) Methane DI only Fig.9 Influence of excess air ratio on heat flux Localheat flux W/m 1 8 - λ=1. 1 3 5 7 Tim e after ignition ms Q C. It is possible to evaluate the tendency of cooling loss in conditions without large change in combustion efficiency. Fig.7 shows a result of the evaluation. Excess air ratios are λ=1. and 1.5. In both excess air ratios, maximum heat release rate increases and combustion period decreases with increase in premixed hydrogen. These mean combustion promotion by hydrogen premixing. On the other hand, maximum local heat flux increases showing increased cooling loss. However, change in Q /Q fuel that is the function of combustion efficiency and cooling loss ratio is small. Premixed hydrogen increases combustion efficiency and cooling loss ratio. Therefore, those two factors are considered to be balanced and the change in Q /Q fuel is small. INFLUENCE OF EXCESS IR RTIO Fig.8 shows influence of overall excess air ratio on combustion characteristics with and without hydrogen premixing. Hydrogen premixing ratio was set at % in LHV. In both conditions heat release rate become sharper with decrease in excess air ratio. Methane direct injection combustion at stoichiometric is slower than that at excess air ratio of 1.. It is considered to be due to over rich condition. t any excess air ratio, combustion with hydrogen premixing has higher and faster heat release than methane direct injection combustion. Instantaneous local heat fluxes in these cases are shown in Fig.9. In both combustion systems, instantaneous local heat flux decreases with increase in excess air ratio. Combustion with hydrogen premixing shows larger heat flux at any condition. Fig.1 shows influence of overall excess air ratio on d Q/ dt max, θ 1-9, Q /Q fuel, and maximum heat flux. Premixed hydrogen reduces combustion period and increases maximum heat release rate. lthough the result means combustion promotion, maximum heat flux is higher than methane at any excess air ratio. Because of the increased cooling loss, Q/Q fuel is lower in conditions of excess air ratio from 1. to.. However, in sufficient lean condition over excess air ratio of around 1.8, Q/Q fuel of hydrogen premixing is higher than methane combustion. Fig.11 shows relation between maximum heat release rate and maximum heat flux with and without hydrogen premixing. The hydrogen premixing ratio was % in LHV. In both combustion systems, maximum heat flux and maximum heat release rate decrease with increase in excess air ratio. However, In the cases with similar level of heat release rate, premixed hydrogen increases maximum heat flux. The difference becomes smaller in lean combustion. Fig.1 shows relation between combustion period and maximum heat flux. With increase in excess air ratio, both combustion period and maximum heat flux increase. In the
cases with similar combustion period, hydrogen premixing shows smaller heat flux than methane direct injection only. These results suggest that although the combustion promotion with hydrogen premixing causes increase in cooling loss, combustion promotion exceeds cooling loss in lean operation conditions. Therefore, the combustion system of methane direct injection stratified combustion in hydrogen lean mixture is effective at lean operation. OBSERVTION OF FLME PROPGTION Fig.13 and Fig.1 show schlieren photography of combustion flame with and without hydrogen premixing. Hydrogen premixing ratio was set at % in LHV. Overall excess air ratio was set at stoichiometric in Fig.13 and 1.5 in Fig.1. In both excess air ratios, combustion is largely promoted with premixed hydrogen to increase flame propagation velocity. Increase in forced convection due to higher flame propagation velocity is considered to increase heat transfer between flame and combustion chamber wall. CONCLUSION Results in this research can be summarized as follows. (1) Premixed hydrogen in methane direct injection stratified charge combustion promotes combustion to shorten combustion periods and to increase maximum value of heat release rate. () Hydrogen premixing tends to increase cooling loss to combustion chamber wall while achieving combustion promotion. (3) Schlieren photography of combustion flame shows increase in flame propagation velocity by premixed hydrogen. () Increase in forced convection due to higher flame propagation velocity of hydrogen is considered to increase heat transfer between combustion gas and combustion chamber wall. (5) In lean operation, combustion promotion with premixed hydrogen exceeds cooling loss to increase Q/Q fuel compared with methane direct injection stratified charge combustion. CKNOWLEDGMENT We would like to thank Miss K. ndo, Mr. K. Shimamura, Mr. T. Futakuchi, and Mr. T. Nagano, former students at Musashi Institute of Technology, for their helps in this research. REFERENCES [1] Shudo, T., Ogawa, H., Miyamoto, N., Combustion and Emissions in a DI Stratified-Charge Engine, Transaction of JSME (in Japanese with English summary), B, Vol., No.57, (199). [] Miyamoto, N., Ogawa, H., Shudo, T., Takeyama, F., Combustion and Emissions in a New Concept DI Stratified Charge Engine with Two-Stage Fuel Injection, SE Technical Paper, No.975, (199). dq /dtmax kj/s θ1-9 ms Q/Qfuel M ax.localheat flux W/m 5 3 1 3 1.8.7..5 H prem ixed+ H /C H =%. 1 8.8 1. 1. 1. 1. 1.8.... Excess air ratio Fig.1 Influence of excess air ratio on combustion Maximum heat release rate kj /s 5 3 1 H prem ixed + H /C H = % lean 8 1 Maximum localheat flux W /m Fig.11 Relationship between maximum local heat flux and maximum rate of heat release θ 1-9 C 3 1 H pre + H /C H = % lean 8 1 Maximum localheat flux W /m Fig.1 Relationship between maximum local heat flux and combustion period 5
. 3.1. 5.3. 7. 8.7ms after spark (a) Methane DI only. 3.1. 5.3. 7. 8.7ms after spark (b) H premixed % + Methane DI Fig.13 Schlieren photography of combustion flame (stoichiometric). 3.1. 5.3. 7. 8.7ms after spark (a) Methane DI only. 3.1. 5.3. 7. 8.7ms after spark (b) H premixed % + Methane DI Fig.1 Schlieren photography of combustion flame (lean λ =1.5) [3] Shudo, T., Ogawa, H., Miyamoto, N., Influence of Fuel Properties on Combustion and Emissions in a DI Stratified Charge Engine, JSE Review, Vol.1, No.1, (1995). [] Ogawa, H., Yamazaki, M., Fujiwara, Y., Tosaka, S., Miyamoto, N., Constant Volume Chamber nalysis for a Two-stage DI Stratified Charge Engine, Proceedings of the 75th JSME Spring nnual Meeting (in Japanese with English summary), Vol.Ⅲ, No.98-1, (1998). [5] ndo, H., Combustion Control Technologies for GDI Engines-Removal of Restriction in Mixing Control-, Mitsubishi Motors Technical Review (in Japanese with English summary),no.1,(1998). [] Lewis, B., Elbe, G., Combustion Flames and Explosion of Gases, (191). [7] Shioji, M., Yamamoto, H., Nishi, T., Ikegami, M., Combustion Characteristics of an Engine Using Gaseous Fuels and Blended Gaseous Fuels, Proceedings of the 1th Internal Combustion Engine Symposium (in Japanese with English summary), (1995). [8] Kido, H., Huang, S., Tanoue, K., Nitta, T., Improving the combustion performance of lean hydrocarbon mixtures by hydrogen addition, JSE Review, Vol.15, No., (1993). [9] Shudo, T., Nakajima, Y., Futakuchi, T., nalysis of Thermal Efficiency in a Hydrogen Premixed Spark Ignition Engine, Proceedings of SME 1999 International Mechanical Engineering Congress, ES-Vol.39, (1999). [1] Shudo, T., Shimamura, K., Nakajima, Y., Combustion and Emissions in a Methane DI Engine with Hydrogen Premixing, JSE Review, Vol.1, No.1, ()