Numerical Study on the Performance Characteristics of Hydrogen Fueled Port Injection Internal Combustion Engine

Transcription

1 American J. o Engineering and Applied Sciences 2 (2): , 2009 ISSN Science Publications Numerical Study on the Perormance Characteristics o Hydrogen Fueled Port Injection Internal Combustion Engine Rosli A. Bakar, Mohammed K. Mohammed and M.M. Rahman Faculty o Mechanical Engineering, Automotive Excellence Center, University Malaysia Pahang, Lebuhraya Tun Razak, Gambang, Kuantan, Pahang, Malaysia Abstract: This study was ocused on the engine perormance o single cylinder hydrogen ueled port injection internal combustion engine. GT-Power was utilized to develop the model or port injection engine. One dimensional gas dynamics was represented the low and heat transer in the components o the engine model. The governing equations were introduced irst, ollowed by the perormance parameters and model description. Air-uel ratio was varied rom stoichiometric limit to a lean limit and the rotational speed varied rom 2500 to 4500 rpm while the injector location was considered ixed in the midway o the intake port. The eects o air uel ratio, crank angle and engine speed are presented in this study. From the acquired results show that the air-uel ratio and engine speed were greatly inluence on the perormance o hydrogen ueled engine. It was shown that decreases the Brake Mean Eective Pressure (BMEP) and brake thermal eiciency with increases o the engine speed and air-uel ratio however the increase the Brake Speciic Fuel Consumption (BSFC) with increases the speed and air-uel ratio. The cylinder temperature increases with increases o engine speed however temperature decreases with increases o air-uel ratio. The pressure luctuations increased substantially with increases o speed at intake port however rise o pressure at the end o the exhaust stroke lead to reverse low into the cylinder past exhaust valve. The luctuation amplitude responded to the engine speed in case o exhaust pressure were given less than the intake pressure. The volumetric eiciency increased with increases o engine speed and equivalent ratio. The volumetric eiciency o the hydrogen engines with port injection is a serious problem and reduces the overall perormance o the engine. This emphasized the ability o retroitting the traditional engines with hydrogen uel with minor modiications. Key words: Hydrogen ueled engine, port injection, air uel ratio, engine speed, cranks angle, perormance characteristics INTRODUCTION In the recent days, there are two main issues regarding the uels: availability and global climate change. The status o the availability o the ossil uels is critical and the prices have been jumped to levels that never been reached beore. Furthermore, the environmental problems are serious and the politics all over the world applied severe conditions or the automotive industry. Researchers, technologists and the automobile manuacturers are increasing their eorts in the implementation o technologies that might be replaced ossil uels as a means o ueling existing vehicles. Hydrogen, as alternative uel, has unique properties give it signiicant advantage over other types o uel. However, the widespread implementation o hydrogen or vehicular application is still waiting several obstacles to be solved. These obstacles are standing in the production, transpiration, storage and utilization o hydrogen. The most important one is the utilization problems. Hydrogen induction techniques play a very dominant and sensitive role in determining the perormance characteristics o the hydrogen ueled internal combustion engine (H 2 ICE) [1]. Hydrogen uel delivery system can be broken down into three main types including the carbureted injection, port injection and direct injection [2]. The port injection uel delivery system (PFI) injects hydrogen directly into the intake maniold at each intake port rather than drawing uel in at a central point. Typically, hydrogen is injected into the maniold Corresponding Author: M. M. Rahman, Automotive Excellence Center, Faculty o Mechanical Engineering, University Malaysia, Pahang, Lebuhraya Tun Razak, Gambang, Kuantan, Pahang, Malaysia 407

2 ater the beginning o the intake stroke. Hydrogen can be introduced in the intake maniold either by continuous or timed injection. The ormer method produces undesirable combustion problems, less lexible and controllable [3]. But the latter method, timed Port Fuel Injection (PFI) is a strong candidate and extensive studies indicated the ability o its adoption [3,4]. The calling sounds or adopting this technique are supported by a considerable set o advantages. It can be easily installed only with simple modiication [5] and its cost is low [6]. The low rate o hydrogen supplied can also be controlled conveniently [7]. External mixture ormation by means o port uel injection also has been demonstrated to result in higher engine eiciencies, extended lean operation, lower cyclic variation and lower NO x production [8,9]. This is the consequence o the higher mixture homogeneity due to longer mixing times or PFI. Furthermore, external mixture ormation provides a greater degree o reedom concerning storage methods. The most serious problem with PFI is the high possibility o pre-ignition and backire, especially with rich mixtures [10-11]. However, conditions with PFI are much less severe and the probability or abnormal combustion is reduced because it imparts a better resistance to backire. Combustion anomalies can be suppressed by accurate control o injection timing and elimination o hot spots on the surace o the combustion as suggested by [5]. Knorr et al. [12] have reported acceptable stoichiometric operation with a bus powered by liquid hydrogen. Their success was achieved by the ollowing measures: Formation o a stratiied charge by timed injection o the hydrogen into the pipes o the intake maniold with a deined pre-storage angle. At the beginning o the intake stroke a rich, non-ignitable mixture passes into the combustion chamber Injection o hydrogen with a relatively low temperature o 0-10 C so that the combustion chamber is cooled by the hydrogen and inally Lowering o the compression ratio to 8:1 Am. J. Engg. & Applied Sci., 2 (2): , 2009 The present contribution introduces a model or a single cylinder, port injection H2ICE. GT-Power sotware code is used to build this model. The main task o this model is to investigate the perormance characteristics o this engine. The emphasis is paid or the trends with the air uel ratio and engine speed. The instantaneous behavior is also considered. MATERIALS AND METHODS One-dimensional basic equations: Engine perormance can be studied by analyzing the mass and energy lows between individual engine components and the heat and work transers within each component. Simulation o one-dimensional low involves the solution o the conservation equations; mass, energy and momentum in the direction o the mean low. Mass conservation is deined as the rate o change in mass within a subsystem which is equal to the sum o m& i and m& e e rom the system dm dt : m& = m& (1) sub i e where, subscript i and e represent the inlet and exit respectively. In one-dimensional low, the mass low rate ( m& ) is expressed as Eq. 2: m& = ρau (2) where, ρ is the density, A is the cross-sectional low area and U is the luid velocity. Energy conservation: The rate o change o energy in a subsystem is equal to the sum o the energy transer o the system. The energy conservation can be written in the ollowing rom: DE DW DQ = + (3) Dt Dt Dt E = The energy W = The work One o the main conclusions drawn rom the Q = The heat experimental study o [10] was the possibility o Energy conservation can be expressed as Eq. 4: overcoming the problem o backire by reducing the injection duration. Sierens and Verhelst [13] DE DW examined DQ } Dt Dt our dierent junctions o the port injection position Dt d(me) dv = p + (uel line) against the air low. Based on the results o m& i H + i m& H h ga(tgas T wall) (4) e e dt dt their CFD model, the junction that gives the highest power output (Y-junction) was dierent rom the junction that gives the highest eiciency (45 junction). e = The internal energy Finally a compromise was suggested. H = The total enthalpy 408

3 h g = The heat transer coeicient T gas and T wall = The temperatures o the gas and wall respectively The heat transer rom the internal luids to the pipe wall is dependent on the heat transer coeicient, the predicted luid temperature and the internal wall temperature. The heat transer coeicient is calculated every time step, which is a unction o luid velocity, thermo-physical properties and the wall surace roughness. The internal wall temperature is given here as input data. Thereore, h g can be expressed as: hg = Cρ UeCp Pr (5) 2 C = The riction coeicient U e = The eective speed outside boundary layer C P = The speciic heat Pr = The Prandtle number The riction coeicient is related to Renolds number which is expressed as Eq. 6: R ρu L v c c e = (6) In case that the wall surace is rough and the low is not laminar, the value o the riction coeicient then is given by Nikuradse s ormula: C (rough) D = The pipe diameter h = The roughness height 0.25 = 1D (2log 10( ) ) 2h 2 (10) Momentum conservation: the net pressure orces and wall shear orces acting on a sub system are equal to the rate o change o momentum in the system: ρu2 & dxa 1 m & & 2 D 2 = dt dx 2 dpa + m i i u + m i eu 4C Cp ρu A (11) u = Fluid velocity D = The equivalence diameter C pl = The pressure loss coeicient Dx = The element length. In order to obtain the correct pressure loss coeicient, an empirical correlation is used to account or pipe curvature and surace roughness, which is expressed as Eq. 12: ρ = The density U c = The characteristic speed L c = The characteristic length and v is the dynamic viscosity by: The riction coeicient or smooth walls is given 16 C =, where ReD < 2000 (7) Re D 0.08 C =, where ReD > 4000 (8) Re 0.25 D 1 2 pi 2 0.5ρu1 The Prandtle number is expressed as Eq. 9: This equation can be extended or the present our stroke engine to: η C p µ Pr = = (9) λ a 2pb BMEP = NVd (14) λ = The heat conduction coeicient µ = The kinematic viscosity and a is the thermal P b = The brake power diusivity N = The rotational speed 409 C p 1 = The inlet pressure p 2 = The outlet pressure u 1 = The inlet velocity p = p (12) Engine perormance parameters: The brake mean eective pressure (BMEP) can be deined as the ratio o the brake work per cycle W b to the cylinder volume displaced per cycle V d and expressed as: BMEP W V b = (13) d

4 Fig. 1: Model o single cylinder, our stroke, port injection hydrogen ueled engine Brake eiciency ( η b ) can be deined as the ratio o the brake power P b to the engine uel energy as: η = b b m ( LHV) (15) m& = The uel mass low rate LHV = The lower heating value o hydrogen P The Brake Speciic Fuel Consumption (BSFC) represents the uel low rate m& per unit brake power output P b and can be expressed as Eq. 16 [14] : BSFC m P = & (16) The volumetric eiciency ( η v ) o the engine deines the mass o air supplied through the intake valve during the intake period, m& a, by comparison with a reerence mass, which is that mass required to perectly ill the swept volume under the prevailing atmospheric conditions and can be expressed as Eq. 17: b Table 1: Hydrogen ueled engine parameters Engine parameter (Unit) Measure Bore (mm) Stroke (mm) Connecting rod length (mm) Piston pin oset (mm) Displacement (liter) Compression ratio Inlet valve close IVC (CA) Exhaust valve open EVO (CA) Inlet valve open IVO (CA) Exhaust valve close EVC (CA) parameters are used to make the model which is shown in Table 1. It is important to indicate that the intake and exhaust ports o the engine cylinder are modeled geometrically with pipes. Several considerations were made the model more realistic. Firstly, an attribute heat transer multiplier is used to account or bends, roughness and additional surace area and turbulence caused by the valve and stem. Also, the pressure losses in these ports are included in the discharge coeicients calculated or the valves. The in-cylinder heat transer is calculated by a ormula which closely emulates the classical Woschni correlation. Based on this correlation, the heat transer coeicient (h c ) can be expresses as Eq. 18 [15] : h 3.26B p T w c = (18) m& a η v = ρ V where, ρ ai is the inlet air density. ai d (17) B = The bore in meters p = The pressure in kpa T = Temperature in K w = The average cylinder gas velocity in m sec 1 Engine model: A single cylinder, our stroke, port injection hydrogen ueled engine was modeled utilizing The combustion burn rate (X b ) using Wiebe the GT-Power sotware. The injection o hydrogen was unction, can be expressed as Eq. 19 [16] : located in the midway o the intake port. The model o n+ 1 the hydrogen ueled single cylinder our stroke port θ θi Xb = 1 exp[ a ] (19) inject engine is shown in Fig. 1. The speciic engine θ 410

5 θ = The crank angle θ i = The start o combustion θ = Combustion period a and n = Adjustable parameters RESULTS AND DISCUSSION This search is categorized into three subsections. The irst part represents the eects o BMEP, BSFC, Brake eiciency and maximum cylinder temperature with the Air-Fuel Ratio (AFR). The second part demonstrates the instantaneous results i.e. variations o intake, exhaust port and cylinder pressure against the crank angle. The third part presents the eects o engine speed (RPM) on engine perormance. It is worthy to mention that one o the most attractive combustive eatures or hydrogen uel is its wide range o lammability. A lean mixture is one in which the amount o uel is less than stoichiometric mixture. This leads to airly easy to get an engine start. Furthermore, the combustion reaction will be more complete. Additionally, the inal combustion temperature is lower reducing the amount o pollutants. Figure 2 shows the eect o air-uel ration on the brake mean eective pressure. The air-uel ratio AFR was varied rom stoichiometric limit (AFR = 34.33:1 based on mass where the equivalence ratio φ = 1) to a very lean limit (AFR = based on φ = 0.2) and engine speed varied rom rpm. BMEP is a good parameter or comparing engines with regard to design due to its independent on the engine size and speed. I torque used or engine comparison, a large engine was always seem to be better when considering the torque, however, speeds become very important when considered the power [17]. It can be seen that the decreases o the BMEP with increases o AFR and speed. It is obvious that the BMEP alls with a nonlinear behavior rom the richest condition where AFR is to the leanest condition where the AFR is The dierences o BMEP are increases with the increases o speed and AFR. The dierences o the BMEP are decreases bar at speed o 4500 rpm while 6.12 bar at speed 2500 rpm or the same range o AFR. This implied that at lean operating conditions, the engine gives the maximum power (BMEP = bar) at lower speed 2500 rpm) compared with the power (BMEP = 0.18 bar) at speed 4500 rpm. Due to dissociation at high temperatures ollowing combustion, molecular oxygen is present in the burned gases under stoichiometric conditions. Thus some additional uel can be added and partially burned. This increases the temperature and the number o moles o the burned gases in the cylinder. These eects increases the pressure were given increase power and mean eective pressure [15]. Am. J. Engg. & Applied Sci., 2 (2): , 2009 Fig. 2: Variation o brake mean eective pressure with air uel ratio or various engine speeds Fig. 3: Variation o brake thermal eiciency with air uel ratio Figure 3 shows the variation o the brake thermal eiciency with the air uel ratio or the selected speeds. It is seen that the brake power (useul part) as a percentage rom the intake uel energy. The uel energy are also covered the riction losses and heat losses (heat loss to surroundings, exhaust enthalpy and coolant load). Thereore lower values o ηb can be seen in the Fig. 3. It can be observed that the brake thermal eiciency is increases nearby the richest condition (AFR 35) and then decreases with increases o AFR and speed. The operation within a range o AFR rom (φ = ) give the maximum values or ηb or all speeds. Maximum ηb o 31.8% at speed 2500 rpm can be seen compared with 26.8% at speed 4500 rpm. Unaccepted eiciency ηb o 2.88% can be seen at very lean conditions with AFR o (φ = 0.2) or speed o 4500 rpm while the eiciency was observed 20.7% at the same conditions 411

6 with speed o 2500 rpm. Clearly, rotational speed has a major eect in the behavior o ηb with AFR. Higher speeds lead to higher riction losses. Figure 4 show the behavior o the brake speciic uel consumption BSFC with AFR. The AFR or optimum uel consumption at a given load depends on the details o chamber design (including compression ratio) and mixture preparation quality. It varies or a given chamber with the part o throttle load and speed range [15]. It is clearly seen rom Fig. 4 that the higher uel is consumed at higher speeds and AFR due to the greater riction losses that can occur at high speeds. It is easy to perceive rom Fig. 4 that the increases o BSFC with decreases in the rotational speed and increases the value o AFR. However, the required minimum BSFC were occurred within a range o AFR rom (φ = 0.9) (φ = 0.7) or the selected range o speed. At very lean conditions, higher uel consumption can be noticed. Ater AFR o (φ = 0.3) the BSFC goes up rapidly, especially or high speed. At very lean conditions with AFR o (φ = 0.2), a BSFC o g kw-h 1 was observed or the speed o 2500 rpm while g kw-h 1 or speed o 4500 rpm. The value BSFC at speed 2500 rpm was observed around 2 times at speed 4000 rpm however around 7 times at speed 4500 rpm. This is because o very lean operation conditions can lead to unstable combustion and more lost power due to a reduction in the volumetric heating value o the air/hydrogen mixture. This behavior can be more clariied by reerring to Fig. 3, where the brake eiciency reduced considerably at very lean operation conditions. Figure 5 shows how the AFR can aect the maximum temperature inside the cylinder. In general, lower temperatures are required due to the reduction o pollutants. It is clearly demonstrated how the increase in the AFR can decrease the maximum cylinder temperature with a severe steeped curve. The eect o the engine speed on the relationship between maximum cylinder temperatures with AFR seems to be minor. At stoichiometric operating conditions (AFR = 34.33), a maximum cylinder temperature o K was recorded. This temperature dropped down to 1350 K at AFR o (φ = 0.2). This lower temperature inhibits the ormation o NO x pollutants. In act this eature is one o the major motivations toward hydrogen uel. The intake port and exhaust port pressures distributions in terms o crank angle are shown in Figure 6 and 7 respectively. The instantaneous behavior is at the 150th cycle or Wide Open Throttle (WOT) Fig. 5: Variation o maximum cylinder temperature with air uel ratio Fig. 4: Variation o brake speciic uel consumption with air uel ratio or dierent engine speed 412 Fig. 6: Instantaneous intake port pressure distributions with crank angle or dierent speed

7 Fig. 7: Instantaneous exhaust port pressure distributions with crank angle or various engine speeds Fig. 8: Instantaneous cylinder pressure distributions with crank angle or various engine speed and stoichiometric operation. Figure 6 and 7 are very important to investigate the backire or pre-ignition occurrence in details. However, or the present case there is neither backire nor pre-ignition and this is the case o normal combustion and shows typical results o pressure variation. The crank angle axis is divided into our parts to indicate the our strokes which take two cycles (720 ). The pressure seems to be like a series o pulses. Each pulse is approximately sinusoidal in shape due to the single cylinder engine. The complexity o the phenomena that occur is apparent. The amplitude o the pressure luctuations increases substantially with increasing engine speed. From Fig. 6, the maximum intake pressure was recorded bar at speed 4500 rpm during the compression stroke while bar at speed o 2500 rpm. At the suction stroke, when high intake vacuum is occurred, the curve is continuously inward and low pulsation is small. For high speed, larger pulses can be seen. At high speeds more uel is required and consequently more vacuum in the intake port. A vacuum o bar was calculated in 4500 rpm compared with bar at 2500 rpm. The gas dynamic eects play a very important rule here. It distorts the exhaust low which is shown in Fig. 7. The rise o the pressure at the end o the exhaust stroke can lead to reverse low into the cylinder past the exhaust valve, however, the high vacuum in the beginning o the irst stroke is highly desired to banish the burnt gases out o the cylinder. At speed o 3000 rpm, a maximum pressure o bar and maximum vacuum o bar were recorded. The response o luctuation o the amplitude to the engine speed in case o exhaust pressure seems to be less than the intake pressure. 413 Figure 8 shows the behavior o the cylinder pressure at the last cycle (150th cycle) or WOT and stoichiometric operation conditions. The behavior o the pressure ollows the combustion phenomenon that occurs. The eect o the rotational speed on the instantaneous behavior o the cylinder pressure is minor. This curve can be divided into three parts or discussion purpose. The irst part corresponds the lame development period which consumes about 5% o the air uel mixture. Very little pressure rise is noticeable and little or no useul work is produced. The second part corresponds the lame propagation period which consumes about 90% o the mixture. During this time, pressure in the cylinder is greatly increased, providing the orce to produce work in the expansion stroke. The maximum values are 51.6 bar at speed o 4500 rpm and bar at speed o 2500 rpm. These values are less than that o traditional gasoline uel about 70 bar with approximately similar conditions. The third part corresponds to lame termination period which consumes about the rest o the mixture (5%). In general this behavior is like the behavior o the traditional gasoline uel, however, it is necessary to keep in mind that during the hydrogen combustion, the lame velocity is rapid and the main changes o cylinder pressure (the second part) occur in a shorter time. Figure 9 shows the variation o the volumetric eiciency with the engine speed. In general, it is desirable to have maximum volumetric eiciency or engine. The importance o volumetric eiciency is more critical or hydrogen engines because o the hydrogen uel displaces large amount o the incoming air due to its low density ( kg m 3 at 25 C and 1 atm.). This reduces the volumetric eiciency to high extent.

8 Fig. 9: Eect o volumetric eiciency with the rotational speed or dierent equivalence ratio For example, a stoichiometric mixture o hydrogen and air consists o approximately 30% hydrogen by volume, whereas a stoichiometric mixture o ully vaporized gasoline and air consists o approximately 2% gasoline by volume [18]. Thereore, the low volumetric eiciency or hydrogen engine is expected compared with gasoline engine works with the same operating conditions and physical dimension. This lower volumetric eiciency is apparent in Fig. 9. Leaner mixture gives the higher volumetric eiciency. The maximum volumetric eiciency was observed 79.4% at lean conditions with AFR = (φ = 0.2) while 62.4% at stoichiometric conditions. Higher speeds lead to higher volumetric eiciency because o the higher speeds give higher vacuum at the intake port and consequent larger air low rate that goes inside the cylinder. Further increase in engine speed leads toward the maximum value o ηv For the considered speeds and with equivalence ratios o 1, 0.8 and 0.6, the maximum η v was recorded at 4200 rpm. For equivalence ratio o 0.4 and 0.2, the maximum η v was recorded at 3800 rpm. At urther higher engine speeds beyond these values, the low into the engine during at least part o the intake process becomes chocked. Once this occurs, urther increase in speed do not increase the low rate signiicantly so volumetric eiciency decreases sharply. This sharp decrease happens because o higher speed is accompanied by some phenomenon that have negative inluence on η v. These phenomenon include the charge heating in the maniold and higher riction low losses which increase as the square o engine speed. In act a lot o solutions were suggested to solve this problem. Furuhama and Fukuma [19] and Lynch [20] suggested and carried out tests with pressure boosting systems or hydrogen engines. White et al. [18] suggest direct injection (incylinder) or hydrogen. 414 Fig. 10: Variation o combustion duration with engine speed or dierent equivalence ratio Figure 10 shows the combustion duration as a unction o the engine speed or dierent equivalence ratio. As stated earlier, hydrogen combustion velocity (1.85 m sec 1 ) is rapid compared with that o gasoline ( m sec 1 ). Thereore short combustion duration is expected. It is well established that the duration o combustion in crank angle degrees only increases slowly with increasing speed or gasoline and diesel engines [15]. Figure 10 shows that this act is also true or hydrogen engines. The luctuation shown is very small, however, it was enlarged in Fig. 10 with a very high scale. All the changes take place within a range o This is too small value, especially i one knows that at 4500 rpm, the crank shat rotates within 1 sec. CONCLUSION The present study considered the perormance characteristics o single cylinder hydrogen ueled internal combustion engine with hydrogen being injected in the intake port. The emphasis was paid to the eects o engine speed, AFR. The instantaneous behavior was also studied. The ollowing conclusions are drawn: At very lean conditions with low engine speeds, acceptable BMEP can be reached, while it is unacceptable or higher speeds. Lean operation leads to small values o BMEP compared with rich conditions Maximum brake thermal eiciency can be reached at mixture composition in the range o (φ = ) and it decreases dramatically at leaner conditions The desired minimum BSFC occurs within a mixture composition range o (φ = ). The operation with very lean condition (φ<0.2) and high engine speeds (>4500) consumes unacceptable amounts o uel Lean operation conditions results in lower maximum cylinder temperature. A reduction o

9 around 1400 K can be gained i the engine works properly at (φ = 0.2) instead o stoichiometric operation Hydrogen combustion results in moderate pressures in the cylinder. This reduces the compactness required in the construction o the engine. But, i abnormal combustion like preignition or backire happens, higher pressures may destroy the connecting rod and piston rings. Thereore, much care should be paid or this point The low values o volumetric eiciency seem a serious challenge or the hydrogen engine and urther studied are required In general, the behavior o the most studied parameters is similar to that o gasoline engine. This gives a great chance to retroit gasoline engines with hydrogen uel with minor modiications. Further uture experimental work will be done to emphasize this simulation and get more details. ACKNOWLEDGEMENT The researchers would like to express their deep gratitude to University Malaysia Pahang (UMP) or provided the laboratory acilities and inancial support. REFERENCES 1. Suwanchotchoung, N., Perormance o a spark ignition dual-ueled engine using splitinjection timing. Ph.D. Thesis, Vanderbilt University, Mechanical Engineering. 2. Das, L.M., Fuel induction techniques or a hydrogen operated engine. Int. J. Hydro. Energ., 15: DOI: / (90)90020-Y 3. Das, L., R. Gulati and P. Gupta, A comparative evaluation o the perormance characteristics o a spark ignition engine using hydrogen and compressed natural gas as alternative uels. Int. J. Hydro. Energ., 25: DOI: /S (99) Das, L., Hydrogen engine: Research and development programs in Indian Institute o Technology (IIT), Delhi. Int. J. Hydro. Energ., 27: DOI: /S (01) Lee, S.J., H.S. Yi and E.S. Kim, Combustion characteristics o intake port injection type hydrogen ueled engine. Int. J. Hydro. Energ., 20: DOI: / (94) Li, H. and G.A. Karim, Hydrogen ueled spark-ignition engines predictive and experimental perormance. J. Eng. Gas Turbines Power, ASME., 128: DOI: / Sierens, R. and S. Verhelst, Experimental study o a hydrogen-ueled engine. J. Eng. Gas Turbines Power, ASME, 123: DOI: / Yi, H.S., K. Min and E.S. Kim, The optimized mixture ormation or hydrogen uelled. Int. J. Hydro. Energ., 25: DOI: /S (99) Kim, Y.Y., J.T. Lee and J.A. Caton, The development o a dual-injection hydrogen-ueled engine with high power and high eiciency. J. Eng. Gas Turbines Power, ASME., 128: DOI: / Ganesh, R.H., V. Subramanian, V. Balasubramanian, J.M. Mallikarjuna, A. Ramesh and R.P. Sharma, Hydrogen ueled spark ignition engine with electronically controlled maniold injection: An experimental study. Ren. Energ., 33: DOI: /j.renene Kabat, D.M. and J.W. Heel, Durability implications o neat hydrogen under sonic low conditions on pulse-width modulated injectors. Int. J. Hydro. Energ., 27: DOI: /S (02) Knorr, H., W. Held, W. Prümm and H. Rüdiger, The man hydrogen propulsion system or city buses. Int. J. Hydro. Energ., 23: DOI: /S (97) Sierens, R. and S. Verhelst, Inluence o the injection parameters on the eiciency and power output o a hydrogen ueled engine. J. Eng. Gas Turbines Power, ASME., 125: DOI: / Blair, G.P., Design and Simulation o our Stroke Engines. 1st Edn., SAE International Society o Automotive Engineers Inc., Warrandale, Pa., USA., ISBN: , pp: Heywood, J.B., Internal Combustion Engine Fundamentals. 1st Edn., McGraw-Hill, London, ISBN-10: X, pp: Ferguson, C.R. and A.T. Kirkpatrick, International Combustion Engines: Applied Thermosciences. 2nd Edn., John Wiley and Sons, Inc., New York, ISBN: 10: , pp: Pulkrabek, W.W., Engineering Fundamentals o the Internal Combustion Engines. 2nd Edn., Prentic Hall, New York, USA., ISBN: 10: White, C.M., R.R. Steeper and A.E. Lutz, The hydrogen-ueled internal combustion engine: A technical review. Int. J. Hydro. Energ., 31: DOI: /j.ijhydene Furuhama, S. and T. Fukuma, High output power hydrogen engine with high pressure uel injection, hot surace ignition and turbocharging. Int. J. Hydro. Energ., 11: DOI: / (86) Lynch, F.E., Parallel induction: A simple uel control method or hydrogen engines. Int. J. Hydro. Energ., 8: DOI: / (83)