Performance and combustion characteristics of a direct injection SI hydrogen engine

International Journal of Hydrogen Energy 32 (2007) 296–304www.elsevier.com/locate/ijhydene

Performance and combustion characteristics of a direct injectionSI hydrogen engine

Ali Mohammadia,∗, Masahiro Shiojib, Yasuyuki Nakaib, Wataru Ishikurab, Eizo Taboc

aEngine Technology, Powertrain, Toyota Motor Europe, Technical centre Hoge Wei 33 Zaventem 1930, BelgiumbGraduate School of Energy Science, Kyoto University, Yoshida-honmachi Sakyo-ku Kyoto 605-8501, Japan

cEnvironmental and Technical Affairs Department, Mitsubishi Motors Co., 5-33-8 Shiba Minatoku Tokyo 108-8401, Japan

Available online 17 August 2006

Abstract

Hydrogen with low spark-energy requirement, wide flammability range and high burning velocity is an important candidate for being used asfuel in spark-ignition engines. It also offers CO2 and HC free combustion and lean operation resulting in lower NOx emissions. However, wellexamined external mixing of hydrogen with intake air causes backfire and knock especially at higher engine loads. In addition, low heatingvalue per unit of volume of hydrogen limits the maximum output power. In this study, attention was paid to full usage of hydrogen advantageemploying internal mixing method. Hydrogen was directly injected into cylinder of a single-cylinder test engine using a high-pressure gasinjector and effects of injection timing and spark timing on engine performance and NOx emission were investigated under wide engine loads.The results indicate that direct injection of hydrogen prevents backfire, and that high thermal efficiency and output power can be achievedby hydrogen injection during late compression stroke. Moreover, by further optimization of the injection timing for each engine load, NOx

emission can be reduced under the high engine output conditions.� 2006 International Association for Hydrogen Energy. Published by Elsevier Ltd. All rights reserved.

Keywords: Internal combustion engine; Direct-injection SI hydrogen engine; Combustion process; Performance; NOx emission

1. Introduction

Hydrogen has been regarded as a future secondary fuel forpower systems due to CO2 and HC free operation. Recent dras-tic increase in the price of petroleum, rapid increase in emissionof green house gases and very strict environmental legislationsare major motivating factors for usage of hydrogen in fuel cellsand internal combustion engines. Nevertheless, energy policyexperts generally believe that utilization of hydrogen as a fuelfor transportation sectors more likely can be expected in in-ternal combustion engines than in fuel cells at least for somedecades [1,2]. Although, there are still several barriers for uti-lization of hydrogen as fuel in transportation sectors, results ofrecent works on hydrogen production, distribution and storageare promising [3].

Hydrogen fuel exhibits desirable characteristics for the com-bustion in SI engines. Wide range of flammability limits enables

∗ Corresponding author.E-mail address: [email protected] (A. Mohammadi).

0360-3199/$ - see front matter � 2006 International Association for Hydrogen Energy. Published by Elsevier Ltd. All rights reserved.doi:10.1016/j.ijhydene.2006.06.005

a smooth engine operation at a very lean mixture with lowNOx level, and the throttle control is unnecessary even at theidling condition in hydrogen engines. Also, the high burningvelocity of hydrogen may contribute to the relatively high ther-mal efficiency with a shorter combustion period at the ignitiontiming close to top dead center. Additionally, low spark energyrequirement would lower the cyclic combustion fluctuation.These preferable natures have encouraged a number of studiesto establish the high-performance hydrogen engine for variousdesign and operating conditions [4–11]. Nevertheless, the ex-tremely rapid combustion of hydrogen causes abnormal burn-ing such as knock, pre-ignition and backfire at higher loads thatprevents a reliable operation and restricts the engine power.Furthermore, a significantly higher burning velocity leads tolower the thermal efficiency compared with other gaseous fuelsbecause of an increase in heat loss due to the knock-like ex-plosive combustion [12–18]. To overcome such problems withhydrogen engines, recently, internal mixing of hydrogen hasbeen examined by some researchers [19,20]. Generally, inter-nal mixing of hydrogen is achieved by high-pressure injectionsystems. This technique is very effective to prevent the backfire

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A. Mohammadi et al. / International Journal of Hydrogen Energy 32 (2007) 296–304 297

Nomenclature

(dp/d�)max maximum pressure increase rate, MPa/deg(dq/d�)max maximum heat release rate, kJ/kg degNOx concentration of NOx in engine exhaust,

ppmp in-cylinder pressure, MPape brake mean effective pressure, MPapmax maximum in-cylinder pressure, MPa

�pi standard deviation of indicated mean effectivepressure normalized by its mean value

� equivalence ratio�e brake thermal efficiency� crank angle, ◦ATDC�i spark timing, ◦BTDC�j injection timing, ◦BTDC

and knock especially under the high engine loads. However,high heat loss due to the high burning velocity of hydrogen isstill a major barrier to achieve a thermal efficiency as high aswhat seen in the recent advanced direct-injection diesel engines.

Present study aims at exhibiting the feasibility of hydro-gen in SI engine by full usage of hydrogen advantages. Thiswas realized by internal mixing of hydrogen with intake airusing in-cylinder high-pressure injection method in a singlecylinder SI engine. The results of experiment indicate that byproper selection of injection timing and spark timing great pro-motion of engine performance can be achieved under variousengine loads.

2. Experimental

The test engine used was a single-cylinder four-cycle naturalaspirated SI engine with a bore diameter of 102 mm and a strokeof 105 mm, which was converted from a direct-injection dieselengine (Yanmar NFD-170). Fig. 1 shows the configurationof the combustion chamber, together with injector and sparkplug arrangements. In this study, flat-head piston was used toform the disc-shaped combustion space, and compression ratiowas fixed at 11.5:1. Hydrogen gas from a high-pressure vesselwas supplied to an electro-magnetically actuated gas injector(Westport Innovation Inc.), and hydrogen flow rate was pre-ciously measured by a mass flow controller (Oval, F-123S).The gas injector has seven hole nozzles with a hole diameter of0.52 mm. Injection timing and duration can be adjusted by anelectronic control system (Morigiken). Injection pressure wasregulated at 8 MPa, which might ensure the sufficient penetra-tion of fuel jet and afford a fast fuel–air mixing even at injectiontiming near top dead center. A passenger-car spark-plug (NGKC7HA) was installed on the cylinder head as ignition sourceand ignition timing was preciously detected by a current mon-itor (Pearson 110). In present study, the distance between thespark plug and injection nozzle tip is approximately 30 mm.This distance is the minimum distance that can be achieved dueto the cylinder head geometry. Engine specifications are listedin Table 1. Fig. 2 shows the schematic of experimental setupfor controlling engine speed, ignition timing and fuel–air ratio,and for measuring engine performance, combustion processand exhaust emissions. In-cylinder pressures were monitoredusing a piezoelectric transducer (Kistler 6052A). In-cylinderpressure data were used to calculate the heat release rate basedon the first law of thermodynamics considering temperaturedependency of specific heat of mixture and the heat loss

Fig. 1. Combustion chamber geometry and gas jet arrangement.

Table 1Engine specifications

Engine type spark-ignition 4-stroke cycleBore × Stroke 102 mm × 105 mmDisplacement 857 ccCompression ratio 11.5:1Combustion chamber Disc shapeSwirl ratio 2.6Intake valve open 360◦ ATDC

close 580◦ ATDCExh valve open 130◦ ATDC

close 380◦ ATDCInjector nozzle 0.52 mm × 7

298 A. Mohammadi et al. / International Journal of Hydrogen Energy 32 (2007) 296–304

P

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Fig. 2. Experimental setup.

to the combustion chamber wall. NOx (NO + NO2) emis-sion was measured using a chemiluminescence NOx analyzer(Yanaco KA-200). All experiments were conducted at wide-open throttle condition, at a fixed cooling-water temperature of80 ◦C, and at a fixed engine speed of 1200 rpm. In present study,beside knocking and unstable combustion, two other limitingcriteria were considered for the engine operation. Engine opera-tion was carried out for exhaust gas temperature and maximumin-cylinder pressure lower than 600 ◦C and 7 MPa, respectively.In-cylinder pressure higher than 7 MPa would damage the gasinjector, when the injection pressure is set at 8 MPa.

3. Results and discussions

In the experiment, effects of fuel–air mixing on engine per-formance and emission were investigated varying the injectiontiming in a very wide range. At first, hydrogen was injectedat early stage of the intake stroke to achieve sufficient mix-ture preparation period. Then, injection timing was delayedtoward the compression stroke and improvement in engineperformance was demonstrated. In the next step, injection wasfurther delayed and optimized injection timing for each engineload was clarified.

3.1. Hydrogen injection during intake stroke

In the first stage of the experiment, hydrogen was injectedinto the cylinder at �j = 300◦ BTDC, when intake valve wasopen during the entire injection period. Fig. 3 shows the effectsof spark timing �i on engine performance, combustion charac-teristics and emission when equivalence ratio � was varied inthe range of 0.3–0.7. At � = 0.5, amount of hydrogen injected

in one stroke is approximately 11.2 mg (0.125NL) with approx-imately 60◦ CA injection duration. As shown, maximum brakethermal efficiency �e is achieved at � = 0.5 and is about 35%.Spark timing �i for maximum brake mean effective pressurepe (MBT) delays with increasing the equivalence ratio �. For� = 0.6 and 0.7, advancing �i toward MBT was difficult dueto strong knocking. Therefore, �e was diminished when com-pared with that for � = 0.5. Brake mean effective pressure peincreases with increasing the equivalence ratio. However, thistendency decays under the high equivalence ratio conditionsand therefore maximum mean effective pressure pe achieved inthis experiment was about 0.65 MPa, which is lower than thatachieved from the original diesel engine. As shown later in thispaper, in this injection timing, increasing amount of hydrogeninjected lowers the amount of the intake air and therefore vol-umetric efficiency decreases. In particular, for hydrogen withlow heating value per unit volume, maximum power is greatlylimited by reduction in the intake air amount. Regarding com-bustion characteristics, maximum in-cylinder pressure pmax de-creases when spark �i is retarded, and is lower than 6 MPa inall conditions. Maximum pressure increase rate (dp/d�)max in-creases by increasing the equivalence ratio �, resulting in anincrease in NOx emissions, typically at � > 0.5. For � = 0.7,maximum NOx emission was about 8000 ppm.

3.2. Hydrogen injection during compression stroke

In the second stage of the experiment, hydrogen was injectedinto the cylinder at the early stage of compression stroke of�j = 130◦ BTDC, just after intake valve close at 140◦ BTDC.Fig. 4 indicates the engine performance, combustion character-istics and NOx emission, when � was varied. At �=0.5, amount


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Fig. 3. Effects of ignition timing �i on brake thermal efficiency �e, brakemean effective pressure pe, maximum pressure increase rate (dp/d�)max,maximum in-cylinder pressure pmax and NOx emission for injection timingof �j = 300◦ BTDC under various equivalence ratios.

of hydrogen injected in one stroke is approximately 13.5 mg(0.152NL) with injection duration of 69◦ CA. Retarding injec-tion from the intake stroke to the early compression stroke re-sults in an increase in pe and �e for each equivalence ratio.Brake mean effective pressure pe increases almost linearly with� and the final pe achieved was about 0.97 MPa, which is higherthan that can be achieved from the original diesel engine. Asshown in Fig. 5, for injection timing of �j = 130◦ BTDC, volu-metric efficiency is slightly affected by increase in the amountof injected fuel. Whereas, for �j =300◦ BTDC drastic reductionin volumetric efficiency can be seen by increase in equivalenceratio �. Here volumetric efficiency is measured intake air flowrate divided by flow rate displaced by the piston (calculated)under standard ambient condition.

Regarding combustion characteristics shown in Fig. 4, retard-ing injection timing causes an increase in pmax and (dp/d�)maxeven under the same equivalence ratio. This is due to increasein amount of fuel injected for a given �. For each equivalenceratio, there is no remarkable changes in NOx emissions when

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Fig. 5. Effects of injection timing �j on volumetric efficiency for differentequivalence ratios.


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compared to that for �j = 300◦ BTDC. This means that de-gree of mixing is not much affected with retard in the injectiontiming.

Finally, injection timing was further retarded to �j =100◦ BTDC and the results are shown in Fig. 6. Comparedwith results of �j = 130◦ BTDC, pe and �e have further in-creased. In this case, spark timing of MBT is not so retarded.Maximum in-cylinder pressure pmax changes little with thatof �j = 130◦ BTDC at the same �i, whereas (dp/d�)max is re-markably increased. Retarding the injection timing lowers NOx

concentration typically for the higher equivalence ratios. Thisindicates that NOx formation depends on the heterogeneity inthe fuel–air mixture. Study carried out by one of the authorsshows that for � < 0.7, combustion of heterogeneous mixturegives higher NOx emission, however, for � > 0.7 NOx emis-sion of homogeneous mixture’s combustion is much higherthan that for heterogeneous mixture [21].

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Fig. 7. Effects of ignition timing �i on brake thermal efficiency �e, maximumpressure increase rate (dp/d�)max, maximum in-cylinder pressure pmax andNOx emissions for � = 0.5 under various injection timings �j.

3.3. Optimization of injection timing for � = 0.5

As shown above, retarding injection timing offers improve-ment in the engine performance and emissions. In the nextstep, the authors tried to optimize the injection timing for� = 0.5, which gave the maximum thermal efficiency. Fig. 7indicates the engine performance, combustion characteristicsand NOx emission when injection timing was varied in therange of �j =130.70◦ BTDC. Retarding the injection timing till�j = 80◦ BTDC increases the thermal efficiency �e and engineoutput pe. However, further injection retard to �j = 70◦ BTDCshows negative effects. For this equivalence ratio, spark timingof �i=5.8◦ BTDC gives the maximum brake thermal efficiencyfor all injection timings. It is clearly shown that at every in-jection timing �j, (dp/d�)max and pmax are decreased with theretarded ignition. Also, at a fixed ignition timing �i, retardingthe injection timing �j increases (dp/d�)max and slightly pmax.However, these values are in the acceptable ranges consideringthe safety of gas injector and combustion noise. NOx emissionlinearly decreases with �i, but its change with �j is not simpleand similar to that of �e. Such results of engine performance arereflected on the combustion processes. Fig. 8 demonstrates thecourses of in-cylinder pressure p and heat-release rate dq/d�measured under the conditions in Fig. 7. Every curve of dq/d�rises just after ignition even when ignition is made after TDC.


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Fig. 8. Changes of in-cylinder pressure p, heat release rate dq/d� with ignition timing �i for different injection timing �j at � = 0.5.

For each �j, the rapid combustion is observed around MBTignition timing, shown by the thick lines.

3.4. Optimization of injection timing under variousequivalence ratios

Injection timing reflects the mixture formation in the vicin-ity of spark plug, then ignition should be adjusted at the propertime that ensures the strong flame kernel and the stable combus-tion followed. Such condition leads to the optimal engine per-formance. As was shown above, in the experimental conditionof the present study, injection timing for best engine operationwas �j =80◦ BTDC for �=0.5. Next, the injection timing wasoptimized under MBT spark timing when equivalence ratio wasvaried in the range of � = 0.3.0.6 and the results are shown inFig. 9. For � = 0.3, the injection timing by which maximum�e is achieved is about 100◦ BTDC and further injection retardseems not effective due to an unstable engine operation. In-creasing � to 0.4 exhibits a great increase in thermal efficiency�e under retard injection timing of �j =90◦ BTDC. For �=0.5in which maximum thermal efficiency �e was achieved, the bestinjection timing is �j = 80◦ BTDC. Although further increasein amount of fuel injection up to �=0.6 offers a higher engineoutput, the thermal efficiency �e is lower than that for � = 0.5.

Although ignition timing is restricted with �j, the stable oper-ation can be obtained in wide range of � as long as the ignition

timing is adequately selected. In Fig. 10, measured values ofbrake thermal efficiency �e, MBT ignition timing �i, relativeintensity of fluctuation of indicated mean effective pressure�pi and NOx concentration in the exhaust are plotted against� for various injection timing �j. Injection at �j = 130◦ and100◦ BTDC particularly exhibits a wider operational range ofequivalence ratio �, whereas the injection retarding to 80◦ and70◦ BTDC restricts the stable combustion to �=0.4.0.6, whichmay prevents the improvement of engine performance. There-fore, further investigation should be required for altering the jetconfiguration, combustion chamber geometry and gas flow thataffect the mixture formation in the combustion space and thecombustion fluctuation. As NOx results indicates, at � > 0.6,NOx emissions can be reduced by retarding the injection tim-ing due to increase in heterogeneity of fuel–air mixture in thisrange of equivalence ratios [21].

Combustion processes at those MBT conditions are shown inFig. 11. At every injection timing �j, combustion proceeds morerapidly as � increases, thus MBT ignition timing is retarded.In particular, in the condition of � greater than 0.7, ignitionshould be made after TDC, through which maximum pressurepmax is decreased and observed at the later crank angle.

Fig. 12 shows the engine performance, combustion char-acteristics and NOx emissions against brake mean effectivepressure pe when injection timing was varied in the range of�j = 70.300◦ BTDC. For injection timing of �j = 300◦ BTDC,


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Fig. 9. Effects of injection timing �j on brake thermal efficiency �e, maximumpressure rise rate (dp/d�)max, MBT injection timing �i and NOx emissionfor different equivalence ratios.

when fuel injection is carried out during the intake stroke,maximum engine output is limited to pe = 0.65. As describedalready, this is due to a decrease in volumetric efficiency,which finally results in knock. Injection timing of 100 and130◦ BTDC extends the engine operation toward lower andhigher pe, where 130◦ BTDC gives slightly higher thermalefficiency at pe = 0.7.0.8. Further delay in injection timing to80◦ BTDC improves the thermal efficiency �e at pe =0.55.0.8.However, in this condition, engine operation under low en-gine load was difficult due to an increase in the combustionfluctuation. In this study, maximum thermal efficiency �e wasachieved for 80◦ BTDC that is 38.9%. As can be seen in theNOx results, NOx emissions for 80◦ BTDC at pe = 0.8 MPais 60% lower than that for 130◦ BTDC. This would be due toincrease in heterogeneity of fuel–air mixture caused by retardin the injection timing [21]. Above results reveal that opti-mization of injection timing is an effective method to obtainhigh thermal efficiency and low NOx emissions for a givenengine load.

4. Summary and conclusions

A direct-injection spark ignition hydrogen engine was de-veloped and attention was paid on the effects of injectiontiming on the engine performance, combustion characteris-tics and NOx emission under a wide range of engine loads.

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emissions against equivalence ratio � for different injection timings �j.

The results reveal that:

1. In-cylinder injection of hydrogen during the intake strokeprevents backfire. However, thermal efficiency and outputpower are limited by knock due to reduction in volumetricefficiency.

2. Hydrogen injection at compression stroke prevents knockand gives an increase in thermal efficiency and maximumoutput power.

3. Hydrogen injection at later stage of compression stroke canachieve the thermal efficiency higher than 38.9% and thebrake mean effective pressure 0.95 MPa. Under high engineoutput conditions, late injection of hydrogen offers a greatreduction in NOx emission due to the lean operation.

As indicated in the present study, employing direct-injectiontechnology in a hydrogen engine is very effective to controlthe abnormal combustion of hydrogen and achieve high ther-mal efficiency and output power. However, the authors believethat further investigation is required for better performance. Al-though late injection results in lower NOx emissions, utilizationof other techniques such as exhaust gas recirculation and after-treatment methods are required to bring the NOx emission toacceptable level. In addition, optimization of combustion cham-ber geometry; injection parameters such as injection pressureand nozzle hole numbers/arrangement; swirl intensity, etc. are


Fig. 11. Changes of in-cylinder pressure p, heat release rate dq/d� with equivalence ratio � for different injection timing �j at MBT ignition timing.

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Fig. 12. Changes of brake thermal efficiency �e, maximum pressure increaserate (dp/d�)max, NOx emission against brake mean effective pressure pe fordifferent injection timings �j at MBT ignition timing.

indeed important to achieve an engine performance level com-petitive to that in the modern direct-injection diesel engines.

Acknowledgment

This study was partially supported by a Grant-in-aid forthe 21st Century COE Program “Establishment of COE onSustainable-Energy System” from Japanese Ministry of Educa-tion, Culture, Sports, Science and Technology and by IndustrialTechnology Research Grant Program in 2003 from New Energyand Industrial Development Organization NEDO of Japan.

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Documents

Performance and combustion characteristics of a direct injection SI hydrogen engine