Study on Combustion and Emission Characteristics of a Combined

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Study on combustion and emission characteristics of a combined injection engine with hydrogen direct-injection Xiumin Yu, Xiongyinan Zuo, Haiming Wu, Yaodong Du, Yao Sun, and Ye Wang Energy Fuels, Just Accepted Manuscript • DOI: 10.1021/acs.energyfuels.6b03300 • Publication Date (Web): 03 Apr 2017 Downloaded from http://pubs.acs.org on April 4, 2017

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Study on combustion and emission characteristics of a combined injection engine with hydrogen directinjection Xiumin Yu, Xiongyinan Zuo*, Haiming Wu, Yaodong Du, Yao Sun, Ye Wang State Key Laboratory of Automotive Simulation and Control, Jilin University, Changchun 130022, China

ABSTRACT:

An

engine

experimental

platform

with

gasoline

port

injection

and

hydrogen/gasoline direct injection was built to study the effects of combined injection mode and ignition timing on combustion and emission characteristics of a spark ignition engine. The study was performed in three modes, including gasoline port injection (GPI) mode, gasoline port injection plus gasoline direct injection (GPI+GDI) mode and gasoline port injection plus hydrogen direct injection (GPI+HDI) mode under the condition of constant excess air ratio and speed. Through the comparison and analysis of the three modes, the results reveal that: combined injection is beneficial to increasing brake power and decreasing brake specific fuel consumption, HC and CO emissions; combustion process will be accelerated and concentrated with combined injection as well as hydrogen addition, and the effects of adding hydrogen is stronger than that of gasoline direct injection; brake thermal efficiency and brake power increase and HC and CO emissions decrease, but NOx emissions increase by adding hydrogen; In any mode, brake power

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and brake thermal efficiency first rises then falls as ignition advance angle increases. HC emissions and NOx emissions decrease by retarding ignition.

1 INTRODUCTION With the energy and environmental issues becoming increasingly prominent, energy saving and emissions reduction have become an inevitable trend in the development of internal combustion engines. In order to meet the requirements above, some researchers focus their efforts on the technology of traditional engines, including VVT (Variable Valve Timing), EGR (Exhaust Gas Recirculation), FSI (Fuel Stratified Injection) combined injection and so on. Among them, combined injection is one of the most effective methods. As the name suggests, this technology is rooted in the cylinder direct injection engine, adding a set of extra fuel injection device to the intake of the engine, so as to achieve the combination of the port fuel injection and cylinder direct injection. Mou et al.1 made experimental studies on a directinjection gasoline engine which was equipped with the combined injection system. The results depicted the low pressure nozzle on the intake manifold can achieve quality homogeneous lean mixture and the direct injection in the cylinder can realize the dense mixture gas near the spark plug. They held that the combined injection makes the lean burn limit of the engine increased. Huang et al.2, 3 did a series of investigations in an ethanol direct injection plus gasoline port injection (EDI+GPI) engine. Their investigations showed that compared with gasoline port injection (GPI) only, EDI+GPI demonstrates stronger effect on charge cooling with lower incylinder temperature and pressure during compression stroke. They also investigated the effect of EDI timing and found that the HC emissions increased at late EDI timing, NO emissions decreased and CO increased with the retarding of EDI timing.

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As mentioned above, in order to alleviate the current energy crisis and environmental problems, other researchers are constantly searching for alternative fuels. As clean and renewable alternative fuel, hydrogen is considered to be the most promising candidate for sparkignition engines4, 5. Compared with conventional hydrocarbon fuels, hydrogen has the merit of low ignition energy, high flame propagation speed, wide flammability and fast diffusion rate6. Jorge Antunes et al.7 investigated the effect of direct injection hydrogen-fuelled mode in a diesel engine. Experimental results indicated that when using hydrogen as a replacement for diesel fuel, the hydrogen-fuelled engine achieved a fuel efficiency of approximately 43% compared with 28% in the conventional, diesel-fuelled mode. They also observed that nitrogen oxides emissions decreased about 20%. C Ji et al.8 investigated the performance of a pure hydrogen-fuelled SI (spark ignition) engine at idle and lean conditions. They discovered both pumping and cooling losses were reduced and NOx and CO emissions were negligible when the engine was leaned out. Ali Mohammadi et al.9 confirmed that using direct-injection technology in a hydrogen engine is very efficient to curb the abnormal combustion of hydrogen and acquire high thermal efficiency and output power. As a result, we can see that there is a potential future in the hydrogen engine. However, in terms of the current development situation, the lack of hydrogen filling infrastructure, low power output and abnormal combustion limit the prospect of hydrogen engine10, 11. Thus, it needs a distance from the mass production of hydrogen engine. In view of the good physical and chemical properties of hydrogen, not only as the single fuel of the engine, but also as blended fuel can play a good role in improving the engine due to the fast burning, high hydrogen-to-carbon ratio, wide flammability and low ignition energy characteristics12, 13, 14. In addition, hydrogen can be obtained by means of onboard hydrogen

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producer and does not require too much modification to the engine, letting the hydrogen mixing engine has great practical significance. In general, most of the studies are derived from engines with port fuel injection (PFI) only, where hydrogen will be injected and premixed. Hamdan et al.15 studied the behavior of compression ignition engine and obtained that thermal efficiency of the engine increases as hydrogen content increases, especially at low engine speed and part-load. According to the experimental results, Yasin et al.16 argued that adding hydrogen to the diesel engine can reduce the smoke emission. Reyes et al.17 found that increasing the hydrogen content in the mixture with natural gas increases its burning velocity, reduces the relative dispersion of combustion and decreases the combustion variability. Gaurav Verma et al.18 made an experiment of adding hydrogen to methane and detected that the thermal efficiency and the peak cylinder pressure increased with increasing H/C ratio for a fixed relative air-fuel ratio (RAFR). Raviteja et al.19 investigated the effect of hydrogen addition on an SI engine fueled with butanol blends. The results indicated that the efficiency of engine improved and an average of 60% reduction in HC and CO emissions was observed upon hydrogen enrichment. Overall, the major ideas of investigating on port hydrogen injection are, higher thermal efficiency18, 19, 20, 23, lower HC raw emissions19,

21, 23

and cycle-to-cycle variations17, 22, shorter combustion duration and ignition

delay17, 20, 22 and higher NOx emissions23. The investigation on hydrogen direct-injection is less comprehensive than that on hydrogen port-injection. Hydrogen with direct injection at compression stroke does not occupy the volume of cylinder for air, effectively avoiding the backfire and increasing the stability of the engine when compared with hydrogen port injection. Besides, the injection quantity, injection timing and injection area can be accurately controlled via direct injection. Yu et al.24, 25, 26 conducted a

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series of experiments on a SI engine with hydrogen direct-injection. On the basis of their experimental results, they considered that the combustion speed increased, mean effective pressure and thermal efficiency was improved with the increasing hydrogen addition fraction. They also found that CO and HC emissions decreased when adding hydrogen by means of direct injection in cylinder. Sato et al.27 developed hydrogen-internal combustion engine system for heavy-duty trucks and confirmed the high-pressure direct-injection combustion system shows that high thermal efficiency and high specific output. Cylinder direct-injection is an efficient hydrogen supply mode and it has become an important research direction due to its high thermal efficiency, high volumetric power and the ability of avoiding backfire28. Through the introduction above, it is undeniable that the combined injection and hydrogen mixing can have a great improvement on the performance of the engine. However, there is limited published literature which has studied the effect of combination of combined injection and hydrogen mixing at present33, 34 and research about combinations of two types of fuel and injection modes is much less. Therefore, three modes were investigated at different ignition timings in this paper, which are gasoline port injection (GPI) mode, gasoline port injection plus gasoline direct injection (GPI+GDI) mode and gasoline port injection plus hydrogen direct injection (GPI+HDI) mode, respectively. GPI+GDI mode and GPI+HDI mode can be considered to be the use of combined injection while GPI+HDI mode can be considered as the combination of combined injection and hydrogen mixing. The experimental results of three modes were compared and analyzed. 2 EXPEROMENTAL SETUP AND PROCEDURE 2.1 Experimental Setup

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The technical specifications of the test engine are listed in Table 1. The schematic diagram of the experimental setup is shown in Fig. 1. The gasoline used in the experiment is commercial gasoline with octane number of 97. The volume percentage of the hydrogen employed in this study is 99.995%. The stoichiometric air-fuel ratios of hydrogen and gasoline are 34.3 and 14.7, respectively. The lower heating value (LHV) of hydrogen and gasoline are 120MJ/kg and 46 MJ/kg, respectively. The engine was connected with an eddy current dynamometer. Measurement variable and measuring instruments needed in the experiment are given in Table 2. Table 1. Technical specifications of the engine Engine type

Four cylinder, four stroke, water cooled, PFI+GDI

Displacement

1.798L

Bore

82.5mm

Stroke

84.2mm

Compression ratio

9.6:1

Rated power/speed

118kW/(5000r/min-6200r/min)

Rated torque/speed

250N·m/(1500r/min-4200r/min)

Valve number /Cylinder

4

Table 2. Measurement deviation of each instrument Parameter

Resolution

Instrument manufacturer

Type

Engine speed

≤±1 r/mim

Luoyang Nanfeng

CW160

Brake power

≤±0.4 N·m

Luoyang Nanfeng

CW160

Gasoline mass flow rate

≤±0.01 g/s

ONO SOKKI(Onokazu detector)

DF-2420

Hydrogen volumetric flow rate

≤±0.2%

Beijing SINCERITY

DMF-1-1AB

Excess air ratio

≤±1.5%

ETAS Engineering TOOLS

LAMBDA LA4

Cylinder pressure

≤±0.5%

ONO SOKKI(Onokazu detector)

DS 9028

Crank angle position

≤±0.01°CA

ONO SOKKI(Onokazu detector)

DS 9028

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Heat release rate

≤±1%

ONO SOKKI(Onokazu detector)

DS 9028

Cycle-to-cycle variation

≤±1%

ONO SOKKI(Onokazu detector)

DS 9028

CO emissions

≤±0.01%

AVL Company

DICOM 4000

HC emissions

≤±1 ppm

AVL Company

DICOM 4000

NOx emissions

≤±1 ppm

AVL Company

DICOM 4000

Figure 1. Schematic diagram of the experimental setup 2.2 Experimental Procedure The combustion and emission performances of three modes were studied in the experiment under the conditions of the throttle opening and engine speed were fixed at 15% and 2000 r/min. The injection timing of cylinder direct injector and port fuel injector were 120°CA BTDC (before top dead center) and 300°CA BTDC and the injection pressure of cylinder direct injector and port fuel injector were 5MPa and 0.3MPa. The experimental procedure can be concisely summarized as follows:

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1. The engine speed was stable at 2000r/min by adjusting the dynamometer. Throttle opening was fixed at 15% and excess air ratio was stable at 1.0. Related data and emissions would be recorded. Each experimental point was recorded for 400 cycles. 2. Reduced the injection pulse width of intake injector to 80% of the original gasoline injection quantity and increased the gasoline injection pulse width of cylinder injector until excess air ratio stabilized at 1.0. Then recorded the related data and emissions. 3. Changed the injection of the cylinder injector from gasoline to hydrogen. Kept the injection pulse width of intake injector unchanged and increased the hydrogen injection pulse width of cylinder injector until excess air ratio stabilized at 1.0. 4. Step 1-3 were repeated at different ignition timing, which are 10°CA BTDC, 15°CA BTDC, 20°CA BTDC, 25°CA BTDC, 30°CA BTDC, respectively. Because the throttle position and excess air ratio were kept fixed throughout the experiment, we can assume that the discharge of fuel in the cylinder direct injector releases 20% of the total heat quantity under GPI+GDI mode. However, due to different properties (such as the stoichiometric air-fuel ratios and LHV) of hydrogen and gasoline, the total heat quantity brought into the cylinder under GPI+HDI mode is 2.3% higher than that under GPI+GDI mode and hydrogen releases 21.8% of the total heat quantity in GPI+HDI mode. 3 RESULTS AND DISCUSSION 3.1 Brake Power and Brake Specific Fuel Consumption (BSFC) 3.1.1 Brake Power Fig. 2 shows comparison of variation of brake power versus ignition timing for GPI mode, GPI+GDI mode and GPI+HDI mode. As it is shown, the trend, brake power first rises and then

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drops with retarding the ignition timing, is the same under three modes. Large ignition advance angle leads to more compression negative work and small ignition advance angle brings relatively more post-combustion, which decreases the power. We can learn from Fig. 2 that the sequence of maximum and average brake power from high to low is: GPI+HDI mode, GPI+GDI mode, GPI mode. As mentioned in experimental procedure, the total heat input in GPI mode is equal to that in GPI+GDI mode and 2.3% less than that in GPI+HDI mode. Comparing GPI mode with GPI+GDI mode, when the total heat input kept unchanged, more brake power outputted by combined injection. Comparing GPI+HDI mode with GPI mode and GPI+GDI mode, the total heat input in GPI+HDI mode is 2.3% higher than that in the other two modes, but the average brake power of GPI+HDI mode is 4.2% higher than that in GPI+GDI mode and 9.2% higher than that in GPI mode. Through the above comparison, a pretty well agreement can be obtained that the brake power can be improved by combined injection. By combined injection, quality homogeneous lean mixture can be formed in intake stroke and local dense mixture near the spark plug can be realized in compression stroke, which is beneficial for forming stratified mixture to make the fuel burn more completely. Therefore, combined injection will increase brake power. However, the reason why brake power of GPI+HDI mode is higher than that in GPI+GDI mode is that the lower ignition energy, higher propagation velocity and shorter quenching distance of hydrogen when compared with gasoline make the fuel burn more completely. Consequently, the brake power of GPI+HDI mode is higher than that of GPI+GDI mode.

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Figure 2. Brake power with respect to ignition timing 3.1.2 Brake Specific Fuel Consumption It can be found from Fig. 3 that BSFC first falls then climbs as the increase of ignition advance angle. For GPI+HDI mode, BSFC is found in gasoline fuel by adding equivalent gasoline fuel amount, based on consumed hydrogen’s LHV, to consumed gasoline fuel. It illustrates that there exists an ignition advance angle which makes the BSFC minimum for a specific engine. Meanwhile, it can be seen that the influence of ignition advance angle on BSFC are sufficiently consistent and even the best ignition advance angle for lowest BSFC is the same in three modes. From Fig. 3, the BSFC of GPI+GDI mode and GPI+HDI mode is lower than that of GPI mode. This may be due to the formation of stratified mixture with combined injection, which promotes the fuel to burn more completely and achieves lower BSFC. In addition, the total input energy in GPI+HDI mode is higher than that of GPI+GDI mode, which will make the theoretical calculation value of BSFC become larger, but the actual BSFC of GPI+HDI mode is still lower than that of GPI+GDI mode. The main reason is the properties of hydrogen such as wide flammability limits, high flame propagation speed and high calorific value32. It makes the combustion process much closer to constant volume combustion and reduces heat transfer loss and exhaust loss24. Moreover, the rate of misfire and incomplete combustion is reduced due to

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hydrogen’s low ignition energy and wild flammability limit. Thus, adding hydrogen can improve brake thermal efficiency.

Figure 3. BSFC with respect to ignition timing 3.2 Combustion Characteristics 3.2.1 Cylinder Pressure Fig. 4 depicts the variation of cylinder pressure with crank angle for GPI+GDI mode and GPI+HDI mode at various ignition timing. It can be found from Fig.4 that the ignition advance angle has a great influence on cylinder pressure. While the influence that the peak cylinder pressure rises and the position of maximum pressure slowly closes to the top dead center with advancing the ignition timing is the same under GPI+GDI mode and GPI+HDI mode. Because the slope of the pressure line is more and more steep with increasing ignition advance angle, we can determine that the pressure rise rate increases by postponing ignition timing. Comparing Fig. 4a and Fig. 4b, the main differences between the two modes are the specific value of peak cylinder pressure and position of maximum pressure, which will be presented in Fig. 5 and Fig. 6. The parameter of peak cylinder pressure (Pmax) has a considerable effect on combustion and emission performance of engines. The variations of peak cylinder pressure with respect to

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ignition timing under the three modes are shown in Fig. 5. The peak cylinder pressure of GPI+GDI mode and GPI+HDI mode is higher than GPI mode at all ignition advance angles under the test. From the data obtained, when compared to GPI+GDI mode, peak cylinder pressure in GPI+HDI mode increases from 2486kPa to 2767kPa (an increase of 11.3% ) at 10°CA BTDC, from 2980kPa to 3318kPa (an increase of 11.3%) at 15°CA BTDC, from 3400kPa to 3734kPa (an increase of 9.8%) at 20°CA BTDC, from 3538kPa to 3924kPa (an increase of 10.9%) at 25°CA BTDC, from 3715kPa to 3983kPa (an increase of 7.2%) at 30°CA BTDC. The increase in peak cylinder pressure could be due to the adiabatic flame propagation speed of hydrogen is higher than other petroleum29, 30. The ideal thermodynamic cycle is more reachable with hydrogen enrichment thanks to the unique combustion properties of hydrogen, such as high flame speed which provides rapid combustion of mixture16. Besides, the short quenching distance of hydrogen helps more fuel burn and work. Therefore, peak gas pressure value increases when adding hydrogen29, 30. Fig. 6 displays the position of maximum pressure with ignition timing for three modes. As we can see, cylinder pressure can be faster to reach the peak in GPI+GDI mode and GPI+HDI mode. When the ignition advance angle is 10°CA BTDC, cylinder pressure almost at the same time reaches the peak under GPI+GDI mode and GPI+HDI mode. However, as ignition advance angle increases, cylinder pressure in GPI+HDI mode is more quickly than that in GPI+GDI mode to reach the maximum pressure and the effect is more pronounced at larger ignition advance angle. Because flame propagation speed and diffusion rate of hydrogen are faster than gasoline and the effect is more pronounced as ignition timing advances when the pressure and temperature are high. In summary, combined injection is favourable to increase peak cylinder pressure and

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advance the position of maximum pressure while adding hydrogen has the same effects and the effects are stronger.

Figure 4. Cylinder pressure with respect to crank angle; a) in GPI+GDI mode; b) in GPI+HDI mode

Figure 5. Pmax with respect to ignition timing

Figure 6. Position of Pmax with respect to ignition timing

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3.2.2 Heat Release Process Fig. 7, Fig. 8 and Fig. 9 describe the instantaneous heat release process versus crank angle for different ignition timings. As we could see, the rule of heat release process is very similar to that of the cylinder pressure. The maximum heat release rate rises and the position of peak heat release rate gradually advances with the advance of ignition. Additionally, retarding ignition will cause the entire combustion process to become longer, which is not conducive to concentrate heat release of fuel. The reason is that the greater ignition advance angle is, the higher piston is near the TDC where mixed gas combustion occurs. Therefore, fuel can be released earlier, which is not only beneficial to the increase of in-cylinder pressure, but also to the rapid release of energy in a short time. For three modes, the maximum heat release of GPI+HDI mode is greater than that of GPI+GDI mode and GPI mode and the heat release process of GPI+HDI mode is faster. This is inspired by the rapid flame propagation speed and high diffusion rate of hydrogen. To conclude, heat release process will be accelerated and concentrated with combined injection as well as adding hydrogen, but the effects of adding hydrogen is more obvious. Through the previous analysis, the average Pmax and maximum heat release rate of GPI+HDI mode is 9.97% and 40.17% higher than that in GPI+GDI mode respectively. The total input energy in GPI+HDI mode is 2.3% higher than that in GPI+GDI mode, which has an effect on the improvement of combustion, but it is not the main reason for such a huge promotion.

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Figure 7. Heat release process with respect to crank angle; a) in GPI+GDI mode; b) in GPI+HDI mode

Figure 8. Maximum heat release rate with respect to ignition timing

Fig. 9. Position of maximum heat release rate with respect to ignition timing 3.3 Emissions Characteristics 3.3.1 Carbon Monoxide (CO) Carbon monoxide emissions versus ignition timing under three modes are given in Fig. 10. The formation of CO has a great relationship with oxygen concentration. For one thing, all experiments in this paper are carried out under the condition that excess air ratio is fixed at 1.0,

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theoretically meaning all the fuel in cylinder will be completely burned; for another, ignition timing changes will not affect the oxygen concentration. Therefore, changing ignition timing has little effect on reduction of the CO emissions. The conclusion is also confirmed through Fig. 10. In either mode, with the ignition timing changing from 10°CA BTDC to 30°CA BTDC, CO emissions change less than 0.2%. From Fig. 10, we also find that CO emissions of GPI+GDI mode are less than that of GPI mode because better formation of mixture in GPI+GDI mode lead to better combustion which contribute to the further oxidation of CO. And CO emissions of GPI+HDI mode are less than that of GPI+GDI mode, because hydrogen does not contain carbon. To sum up, we can assume that combined injection has effect on reducing CO emissions while the effect is pronounced more obviously by adding hydrogen.

Figure 10. CO emissions with respect to ignition timing 3.3.2 Hydrocarbon (HC) HC emissions increase with increasing ignition advance angle for three modes as is evident from Fig. 11. This is because that increasing ignition advance angle will decrease exhaust temperature on the basis of Fig. 12. As a result, the process of oxidation reaction of unburned hydrocarbon will be blocked due to the low temperature, causing the HC emissions rise eventually. It can be seen from Fig. 11 that the HC emissions of GPI+GDI mode and GPI+HDI

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mode are both a lot less than that of GPI mode. This is primarily attributed to the homogeneous lean mixture formed in intake leading a more complete combustion, while the fuel injected from direct injector plays a role of combustion-supporter. It tells us that combined injection is an effective way to reduce HC emissions. From Fig. 12, the exhaust temperature of GPI+GDI mode is slightly higher than that of GPI+HDI mode. The reason why HC emissions of GPI+HDI mode is nearly half of GPI+GDI mode is that the wide flammable limits and fast flame propagation speed of hydrogen is beneficial to combustion completely and hydrogen’s short quenching distance and ability to extend the lean limit can reduce HC emissions31. What’s more, there is no element carbon in hydrogen, this is an important reason for lower HC emissions in GPI+HDI mode. So we have reason to believe that adding hydrogen can greatly reduce HC emissions.

Figure 11. HC emissions with respect to ignition timing

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Figure 12. Exhaust temperature with respect to ignition timing 3.3.3 Nitrogen oxides (NOx) The variations of NOx under three modes are presented in Fig. 13 as ignition advance angle increases. High temperature, rich oxygen, long residence time at high temperature can provide a fantastic environment for the formation of NOx. For a certain mode, the temperature in cylinder rises and the staying timing in high temperature become longer as ignition advance angle increases, which increase NOx emissions. Similarly, the temperature in cylinder increases with combined injection and hydrogen addition, hence the NOx emissions of GPI+GDI mode and GPI+HDI mode are higher than that of GPI mode.

Figure 13. NOx emissions with respect to ignition timing 4 CONCLUSIONS The effects of combined injection mode and ignition timing on combustion and emission of a SI engine were investigated. The experiments were performed under three modes, including GPI mode, GPI+GDI mode and GPI+HDI mode. The conclusions based on experimental results were drawn as follows:

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(1). The combustion and emission performance can be obviously improved through adopting combined injection due to the formation of stratified mixture. By using combined injection, the stratified mixture will burn more completely, which is propitious to increase brake power, brake thermal efficiency and decrease CO and HC emissions. Moreover, combined injection is capable of accelerating and concentrating the combustion process. (2). Adding hydrogen leads to lower BSFC, higher heat release rate, higher brake power, higher combustion temperature and pressure. Hydrogen’s short quenching distance can reduce HC emissions. However, higher temperature brings about the formation of NOx, therefore, NOx emissions will increase with hydrogen addition. (3). Compare GPI+GDI mode and GPI+HDI mode, experimental results display: First, brake power and NOx emissions of GPI+HDI mode is higher than that of GPI+GDI mode. Second, The effects of accelerating of concentrating combustion process in GPI+HDI mode is stronger than that in GPI+GDI mode. Hence, in GPI+HDI mode, the peak cylinder pressure and heat release rate are greater and the position of maximum peak cylinder pressure and heat release rate are more advanced. Third, BSFC, HC emissions and CO emissions in GPI+HDI mode are lower. In summary, NOx emissions will increase after adding hydrogen, but economy, power, combustion and other emission characteristics of the engine are all improved with hydrogen addition. (4). The influence rules of ignition timing on combustion and emission performance are basically consistent under three modes. With the increase of ignition advance angle, the brake power and brake thermal efficiency first rises then falls. And the best ignition advance angle for lowest BSFC is the same in three modes. As for emissions, the hydrocarbon and nitrogen oxides

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emissions decrease by retarding ignition. Furthermore, changing ignition timing has little effect on reduction of the CO emissions. ACKNOWLEDGMENT This work was financially supported by the National Natural Science Foundation of China (Grant No. 51276079), Foundation of State Key Laboratory of Automotive Simulation and Control of Jilin University (Grant No. 20120109), Ph.D. Programs Foundation of Ministry of Education of China (Grant No. 20110061110032), and Project 2016022 supported by Graduate Innovation Fund of Jilin University. The authors appreciate the insightful comments and suggestions from the reviewers and the editor. ABBREVIATIONS GPI, gasoline port injection; GPI+GDI, gasoline port injection plus gasoline direct injection; GPI+HDI, gasoline port injection plus hydrogen direct injection; HC, hydrocarbon; CO, carbon monoxide; NOx, nitrogen oxides; VVT, variable valve timing; EGR, exhaust gas recirculation; FSI, fuel stratified injection; SI, spark ignition; PFI, port fuel injection; LHV, lower heating value; BTDC, before top dead center; BSFC, brake specific fuel consumption; Pmax, peak cylinder pressure. REFERENCES (1) Mou, J. F.; Chen, R. Q.; Lu, Y.W. Applied Mechanics and Materials 2014, 532, 362-366. (2) Huang, Y.; Hong, G. Applied Energy 2015, 160, 244-254. (3) Huang, Y.; Hong, G.; Huang, R. Energy 2016, 111, 92-103. (4) Maghbouli, A.; Yang, W.; An, H.; Shafee, S.; Li, J. Energy 2014, 76, 768-779. (5) Wu, H. W.; Wu, Z.Y. Energy 2012, 47, 411-420. (6) Das, L. M. International Journal of Hydrogen Energy 2002, 27(9), 953-965.

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(7) Antunes, J. M. G.; Mikalsen, R.; Roskilly, A. P. International Journal of Hydrogen Energy 2009, 34(15), 6516-6522. (8) Ji, C. W.; Wang, S. F. International Journal of Hydrogen Energy 2009, 35(1), 346-355. (9) Mohammadi, A.; Shioji, M.; Nakai, Y.; Ishikura, W.; Tabo, E. International Journal of Hydrogen Energy 2007, 32(2), 296-304. (10) Shin, B.; Cho, Y.; Han, D.; Song, S.; Chun, K. M. Fuel and Energy Abstracts 2011, 36(10), 6281-6291. (11) Li, H.; Karim, G. A. International Journal of Hydrogen Energy 2004, 29(8), 859-865. (12) Vudumu, S. K.; Koylu, U. O. Energy 2011, 36(1), 647-655. (13) Zhao, J.; Ma, F.; Xiong, X.; Deng, J.; Wang, L. Energy 2013, 59, 658-665. (14) Deb, M.; Paul, A.; Debroy, D.; Sastry, G. R. K.; Panua, R. S.; Bose, P. K. Energy 2015, 85(2), 658-665. (15) Hamdan, M. O.; Selim, M. Y. E.; Al-Omari, S. A. B.; Elnajjar, E. Renewable Energy 2015, 82(1), 54-60. (16) Yasin, K.; Đlker, G.; Tarkan, S.; Levent, Y.; Ahmet, S. D. International Journal of Hydrogen Energy 2015, 41(1), 656-665. (17) Reyes, M.; Tinaut, F. V.; Melgar, A.; Pérez, A. International Journal of Hydrogen Energy 2016, 41(3), 2064-2074. (18) Verma, G.; Prasad, R. K.; Agarwal, R. A.; Jain, S.; Agarwal, A. K. Fuel 2016, 178, 209-217. (19) Raviteja, S.; Kumar, G. N. International Journal of Hydrogen Energy 2015, 40(30), 95639569. (20) Iorio, S. D.; Sementa, P.; Vaglieco, B. M. International Journal of Hydrogen Energy 2014, 39(39), 9809-9823.

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(21) Kahraman, N.; Çeper, B.; Akansu, S. O.; Aydin, K. International Journal of Hydrogen Energy 2009, 34(2), 1026-1034. (22) Baratta, M.; D'Ambrosio, S.; Misul, D.; Spessa, E. Journal of Engineering for Gas Turbines and Power 2014, 136(5), 775-791. (23) Chen, Y.; Raine, R. Fuel 2016, 176, 190-199. (24) Du, Y. D.; Yu, X. M.; Wang, J. L.; Wu, H. M.; Dong, W.; Gu, J. Q. International Journal of Hydrogen Energy 2016, 41(4), 3240-3248. (25) Niu, R. X.; Yu, X. M.; Du, Y. D.; Xie, H. G.; Wu, H. M.; Sun, Y. Fuel 2016, 186, 792-799. (26) Wu, H. M.; Yu, X. M.; Du, Y. D.; Ji, X. B.; Sun, Y.; Gu, J. Q. Applied Thermal Engineering 2016, 100, 829-839. (27) Sato, Y.; Kawamura, A.; Yanai, T.; Naganuma, K.; Yamane, K. Iasme/wseas International Conference on Energy and Environment 2009, 289-296. (28) Wang, C. Y.; Liu, F. S. Transactions of Csice 2010, 28(6), 519-524. (29) Verhelsta, S.; Woolleyb, R.; Lawesb, M.; Sierensa, R. Proceedings of the Combustion Institute 2005, 30(1), 209-216. (30) Kumar, M. S.; Ramesh, A.; Nagalingam, B. International Journal of Hydrogen Energy 2003, 28(10), 1143-1154. (31) Ji, S.; Wang, S. International Journal of Hydrogen Energy 2010, 35(3), 1453-1462. (32) Quadri, S. A. P.; Masood, M.; Kumar, P. R. Fuel 2015, 157, 279-284. (33) Sun Y.; Yu X. M.; Jiang, L. L. International Journal of Hydrogen Energy 2016, 41(41), 18631-18640. (34) Hamzehloo, A.; Aleiferis, P. Computational Study of Hydrogen Direct Injection for Internal Combustion Engines; SAE Technical Paper 2013-01-2524; SAE International:

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Warrendale, PA, 2013.

Figure 1. Schematic diagram of the experimental setup

Figure 2. Brake power with respect to ignition timing

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Figure 3. BSFC with respect to ignition timing

Figure 4. Cylinder pressure with respect to crank angle; a) in GPI+GDI mode; b) in GPI+HDI mode

Figure 5. Pmax with respect to ignition timing

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Figure 6. Position of Pmax with respect to ignition timing

Figure 7. Heat release process with respect to crank angle; a) in GPI+GDI mode; b) in GPI+HDI mode

Figure 8. Maximum heat release rate with respect to ignition timing

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Fig. 9. Position of maximum heat release rate with respect to ignition timing

Figure 10. CO emissions with respect to ignition timing

Figure 11. HC emissions with respect to ignition timing

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Figure 12. Exhaust temperature with respect to ignition timing

Figure 13. NOx emissions with respect to ignition timing Table 1. Technical specifications of the engine Engine type

Four cylinder, four stroke, water cooled, PFI+GDI

Displacement

1.798L

Bore

82.5mm

Stroke

84.2mm

Compression ratio

9.6:1

Rated power/speed

118kW/(5000r/min-6200r/min)

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Rated torque/speed

250N·m/(1500r/min-4200r/min)

Valve number /Cylinder

4

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Table 2. Measurement deviation of each instrument Parameter

Resolution

Instrument manufacturer

Type

Engine speed

≤±1 r/mim

Luoyang Nanfeng

CW160

Brake power

≤±0.4 N·m

Luoyang Nanfeng

CW160

Gasoline mass flow rate

≤±0.01 g/s

ONO SOKKI(Onokazu detector)

DF-2420

Hydrogen volumetric flow rate

≤±0.2%

Beijing SINCERITY

DMF-1-1AB

Excess air ratio

≤±1.5%

ETAS Engineering TOOLS

LAMBDA LA4

Cylinder pressure

≤±0.5%

ONO SOKKI(Onokazu detector)

DS 9028

Crank angle position

≤±0.01°CA

ONO SOKKI(Onokazu detector)

DS 9028

Heat release rate

≤±1%

ONO SOKKI(Onokazu detector)

DS 9028

Cycle-to-cycle variation

≤±1%

ONO SOKKI(Onokazu detector)

DS 9028

CO emissions

≤±0.01%

AVL Company

DICOM 4000

HC emissions

≤±1 ppm

AVL Company

DICOM 4000

NOx emissions

≤±1 ppm

AVL Company

DICOM 4000

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