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Comparative combustion characteristics of gasoline and hydrogen fuelled ICEs
international journal of hydrogen energy 35 (2010) 5114–5123
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Comparative combustion characteristics of gasoline and hydrogen fuelled ICEs Jonathan Nieminen*, Ninochka D’Souza, Ibrahim Dincer Faculty of Engineering and Applied Science, University of Ontario Institute of Technology, 2000 Simcoe Street North, Oshawa, ON L1H 7K4, Canada
article info
abstract
Article history:
In this paper, some models of comparative combustion characteristics for gasoline and
Received 14 July 2009
hydrogen fuelled spark ignition internal combustion engines were developed and dis-
Received in revised form
cussed from a thermodynamic and heat transfer perspective. The geometry used was that
15 August 2009
of a 3.4L GM V6 engine with a compression ratio of 9.5:1. Models for mass fraction burned,
Accepted 30 August 2009
pressure, temperature, and gas speed were developed according to the literature survey
Available online 2 October 2009
and graphed over the cycle range. Furthermore, Pressure–Volume and Temperature– Entropy models were developed for both gasoline and hydrogen fuelled engines. Analysis
Keywords:
of these models indicated approximately a 6.42% increase thermal efficiency for the
Hydrogen
hydrogen fuelled engine due to less exhaust blow down, less heat rejection during the
Gasoline
exhaust stroke, and its shorter combustion duration closer to TDC. However, it was found
Engines
that the hydrogen fuelled engine had approximately a 35.0% decrease in power output at
Thermodynamic
an equivalence ratio of 1.0 due to the decrease in MEP and a greater amount of heat transfer
Heat Transfer
to the cooling system due to the increased combustion temperatures, shorter quenching distance associated with H2 combustion and greater flame speed. Finally, an increase in cycle temperatures and pressures was observed from increasing the equivalence ratio from 0.4 to 1.0 to 1.2. ª 2009 Professor T. Nejat Veziroglu. Published by Elsevier Ltd. All rights reserved.
1.
Introduction
Due to rising fuel costs, stricter environmental regulations, decreasing oil supplies, and additional problems facing modern society, alternative fuels such as hydrogen have gained popularity. However, that alone is insufficient reason to develop such a technology. The development of a hydrogen internal combustion engine requires a solid understanding of the comparative performance characteristics of the engine. Therefore, comparative pressure models for both gasoline and hydrogen fuelled ICEs are developed as in ‘‘An Analytic Model for Cylinder Pressure in a Four Stroke S.I Engine’’. Since the
pressure and temperature of a gas are related via the Ideal Gas Law; therefore the extension of the pressure model to develop a comparative temperature model for both gasoline and hydrogen fuelling is considered to be justified and accurate. The pressure model will be graphed against volume model in order to make a Pressure–Volume model for both gasoline and hydrogen fuelling. This comparative model will then be used to determine the effect of differing cycle pressure on work output and thermal efficiency between a gasoline and hydrogen fuelled engine. Similarly, the temperature model will be graphed against the entropy generation model in order to make a Temperature–Entropy model for both gasoline and
* Corresponding author. E-mail addresses: [email protected] (J. Nieminen), [email protected] (N. D’Souza), [email protected] (I. Dincer). 0360-3199/$ – see front matter ª 2009 Professor T. Nejat Veziroglu. Published by Elsevier Ltd. All rights reserved. doi:10.1016/j.ijhydene.2009.08.098
international journal of hydrogen energy 35 (2010) 5114–5123
hydrogen fuelling. This comparative model will be used to determine the effect of cycle temperature on thermal efficiency. Also, models for cylinder volume, exposed wall area, and gas speed as a function of crank angle were found [2]. Plentiful background information relating to the thermodynamic analysis of spark ignition internal combustion engines is available and has been used to aid this analysis. In ‘‘An Analytic Model for Cylinder Pressure in a Four Stroke S.I Engine’’ by Eriksson et al. effectively developed a model for cylinder pressure for by interpolating between compression and expansion asymptotes as defined by the Otto Cycle. The reason behind such an interpolation is the result of the heat addition not being isochoric or isobaric. Therefore a model was developed to compensate by the incorporation of the Wiebe Burn Function over the combustion duration in order to interpolate between the polytropic compression and expansion asymptotes. Varying cylinder pressure during combustion is modelled with varying volume as a function of burning rate. The compression asymptote was modelled as a polytropic compression from qbdc to qsoc, and the expansion asymptote was modelled as a polytropic expansion from qeoc to qbdc. The range between qsoc and qeoc is described accurately through the interpolated pressure function. According to Enzo Galloni in ‘‘Combustion Modelling and Performance Estimation of an S.I Engine Using Lean H2-Air Mixtures’’ developed a numerical model that predicted that an ultra-lean, fast burning H2 could be able to control engine load by varying the amount of fuel at wide open throttle (WOT) conditions, thus limiting engine pumping losses. Galloni [4] eluded that this decrease in pumping losses would translate into an increased thermal efficiency. All of this was hypothesized by the model then proven to through experimentation. When compared to gasoline fuelling there are higher cylinder pressures for H2 fuelling exist during unthrottled operation. These higher cylinder pressures were mentioned as a potential safety hazard to due increased mechanical and thermal stresses. ‘‘Hydrogen Internal Combustion Engines’’ by Eiji Tomita covers the principles of H2 fuelling for internal combustion engines. Pressure and knock profiles were constructed for various fuels such as gasoline, n-butane, and hydrogen. It was found that gasoline had the highest cylinder pressures and hydrogen had the lowest. Also, hydrogen and gasoline were found to have similar knock profiles that differ by timing. The hydrogen knock profile was found to reach its maximum value later in the cycle whereas the gasoline knock profile was found to reach its maximum value earlier in the cycle. The latter is more damaging to the engine because it is taking place at higher pressures. According to Szwaja et al. in ‘‘Comparison of Hydrogen and Gasoline Combustion Knock in an Internal Combustion Engine’’ the ignition timing of a hydrogen engine was advanced to approximately 5 to 10 BTDC. Also, they discussed the effects of combustion timing on knock in both hydrogen and gasoline fuelled internal combustion engines. Along with other literature, they had a spark timing of approximately 5 to 10 ATDC and had been found to end around 18 to 20 ATDC. Due to the greater flame speed of hydrogen the spark timing is moved closer to TDC. This yields a combustion duration of approximately 26 to 28 . Finally, a full computer simulation of a hydrogen fuelled, spark ignition ICE was done in ref [15]. Halmari [15] showed the effect of
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retarded and advanced timing on a hydrogen fuelled ICE, simulated engine torque as a function of spark timing, and brake thermal efficiency as a function of equivalence ratio for a supercharged at various compression ratios. It was found that brake torque was a maximum at around 7 BTDC, the brake thermal efficiency peaked at a compression ratio of approximately 12.6:1 and an equivalence ratio of approximately 0.9. With respect to heat transfer within internal combustion engines, much work has been done in the area and was helpful in understanding the phenomenon. Both Kirkpatrick et al. and Heywood in their respective books devote chapters towards the understanding of heat transfer in internal combustion engines, ref. [2] and ref. [3] respectively. According to Shudo et al. in ‘‘Applicability of Heat Transfer Equations to Hydrogen Combustion’’, hydrogen combustion exhibits greater heat loss to the coolant due to higher flame speed and shorter quenching distance. According to Nijeweme et al. in ‘‘Unsteady In Cylinder Heat Transfer in a Spark Ignition Engine: Experiments and Modelling’’ [10], evaluating these temperatures, fluxes, limits, etc not only helps in improving the efficiency of engine performance but it also has a huge impact on exhaust emissions that are exponentially dependant on the peak combustion temperature values. In this paper, comparative combustion characteristics of gasoline and hydrogen fuelled internal combustion engines (ICEs) such as – mass fraction burned (mfb), pressure, temperature, and gas speed. Furthermore, Pressure–Volume and Temperature–Entropy diagrams were developed and commented upon from a thermodynamic and heat transfer perspective.
1.1.
Thermodynamic analysis
To construct a thermodynamic model for the working fluid inside an internal combustion engine, following assumptions are made:
Spark ignition engine operating on the 4-stroke Otto Cycle. Working fluids are ideal gases. The compression and expansion processes are polytropic. Neglecting the effects of friction, valve and piston ring blowby, squish and swirl.
1.2.
Mass fraction burned
The mass fraction burned (mfb) represents the percentage of fuel burned as a function of crank angle throughout the combustion duration. It is modelled by a Weibe Burn Function with m ¼ 2 and a ¼ 5 [2]. q qSOC ; qSOC < q < qEOC f ðqÞ ¼ 1 exp a Dq
1.3.
(1)
Interpolated pressure model
Since combustion in a spark ignition engine takes place over a period of time known as combustion duration it is not isochoric as the Otto cycle suggests. Therefore, development of an accurate model of the behaviour of the pressure during the
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international journal of hydrogen energy 35 (2010) 5114–5123
Fig. 1 – Mass fraction burned (%/100) vs. crank angle (8CA).
combustion process is fundamental to an accurate model of the cycle as a whole [5]. 8 > > > > > > > > > > <
P1
k
vd P1 vðqÞ k k vd vc PðqÞ ¼ ð1 f ðqÞÞP1 vðqÞ þf ðqÞP3 vðqÞ > k > > vc > P3 vðqÞ > > > > k > > vd : P4 vðqÞ
; qTDC q < qBDC ; qBDC q < qSOC ; qSOC q < qEOC
(2)
; qEOC q < qBDC
1.5.
Gas speed in an ICE varies changes with pressure and piston speed. Gas speed is heavily dependent on the pressure rise during combustion. It was modelled as a piecewise function based on whether or not the valves are open or closed and which in cylinder processes are taking place. The in cylinder gas speed can be modelled according to refs [2,3].
; qBDC q < qTDC
1.6. 1.4.
Gas speed
Interpolated temperature model
Since heat release in a real SI engine is not isochoric as the Otto cycle suggests and since temperature and pressure are related via the Ideal Gas Law, therefore the behaviour of the interpolated temperature model is accurate by being a logical extension of the interpolated pressure model. The temperature during the intake stroke was modelled as in refs [2,3] in order to account for the residual fraction of combustion products still in the cylinder after the exhaust stroke. The residual fractions of burned gases were found to be 0.01166 (1.166%) and 0.011826 (1.1826%) for hydrogen and gasoline, respectively. 8 k1 i h k p1 > ; qTDC q < qBDC 1 fr T1 þ fr Texh ðqÞ 1 1 pexh > > ðqÞ > > k1 > > vd > > T1 vðqÞ ; qBDC q < qSOC > > < k1 k1 v v c TðqÞ ¼ ð1 f ðqÞÞT1 d þf ðqÞT3 vðqÞ ; qSOC q < qEOC vðqÞ > > k1 > > > vc > T3 vðqÞ ; qEOC q < qBDC > > > > k1 > : vðqÞ T4 vd ; qBDC q < qTDC
Comparative pressure–volume relationships
A Pressure–Volume (P–V) model is essential to measuring the work output of a thermodynamic gas power cycle. Since gasoline and hydrogen fuelled engines have differing pressure relationships, it follows that the work output of the cycle would be affected. The work done by pressure over the cycle is called the mean effective pressure (MEP). In general, a greater MEP is indicative of a greater work output of the engine. The interpolated pressure model was extended to accommodate exhaust and intake processes. There are 4 processes in a 4stroke Otto cycle [2,3]:
Intake Stroke from qTDC to qBDC Polytropic Compression from qBDC to qTDC Polytropic Expansion from qTDC to qBDC Exhaust Stroke from qBDC to qTDC
1.7.
Comparative temperature – entropy relationships
A Temperature–Entropy (T–s ) model is essential in understanding the amount of heat added or lost during a process and/
(3) where:
fr ¼
1 pexh rc pexp
Table 1 – Combustion durations and ignition timings. !1k (4)
Fuel
qsoc
qeoc
6q
Hydrogen Gasoline
352 330
378 400
26 70
international journal of hydrogen energy 35 (2010) 5114–5123
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Fig. 2 – Cylinder pressure (kPa) vs. crank angle (8CA).
or how irreversible a thermodynamic process is by its deviation from the isentropic lines. The interpolated temperature model was employed here. Furthermore, entropy was found to be a function of the pressure, temperature, and equivalence ratio of the mixture over the entire cycle range [3,14].
law of similarity [12] and proposed a function based on pressure, cylinder bore, gas speed, and temperature [2]. When applied to gasoline ICEs as a function of crank angle it was found to be: hðqÞ ¼ 3:26PðqÞ0:8 UðqÞ0:8 b0:2 TðqÞ0:55
2.
Heat transfer review
Heat transfer within an ICE is directly related to its thermal performance. For the goal and scope of this light analysis, it is reasonable to make a few assumptions. Convection is the dominant mode of heat transfer within the cylinder[2]. Simplistic ‘‘single zone’’ combustion model. Finite heat release method is assumed to be valid.
However, when the aforementioned formula is applied to hydrogen fuelled ICEs it undershoots the amount of heat transfer to the coolant by approximately 75% [11]. To offset this, the convective heat transfer coefficient in Woschni’s correlation was multiplied by 4 to estimate the actual amount of convective heat transfer more accurately. When made a function of crank angle, the Woschni’s correlation took the form of: hðqÞ ¼ 13:04PðqÞ0:8 UðqÞ0:8 b0:2 TðqÞ0:55
2.3. 2.1.
(6)
Other correlations
Comparative heat transfer coefficients
The in cylinder processes change with respect to the crank angle and the nature of the process occurring within the 4stroke cycle. For internal combustion engines, there are several relations for convective heat transfer coefficients that have been calculated through experimentation and/or analysis.
2.2.
(5)
The Nusselt Correlation was constructed from the data on quiescent heat loss measurements from a spherical bomb combustor [13]. It took the form of: 1=3 (7) hðqÞ ¼ 1:151 1 þ 1:24Up PðqÞ2 TðqÞ The Briling Correlation was derived as a slight modification of the mean speed factor term in the Nusselt Correlation [13]. 1=3 hðqÞ ¼ 1:151 3:5 þ 0:185Up PðqÞ2 TðqÞ
Woschni’s correlation
Woschni reviewed various heat transfer correlations for reciprocating engines and proposed a formula based on the
(8)
The Eichelberg Correlation was determined through the measurements of instantaneous heat transfer rates from a 2
Table 2 – Comparative cycle pressures for gasoline and H2 fuelling @ f [ 1.0. Crank Angle ( CA) Gasoline (kPa) Hydrogen (kPa)
qivc ¼ 240
qsoc
qcH2 ¼ 378 qcGAS ¼ 382
qeoc
qbdc ¼ 540
96.451 96.451
843.005 1496.32
7628.1027 6192.6083
5214.3716 6192.6083
590.2246 334.0109
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Fig. 3 – Cylinder temperature (K) vs. crank angle (8CA) @ f [ 1.0.
stroke diesel engine [13]. Through solving the temperature variation through a harmonic analysis, he concluded that: 1=3 1 hðqÞ ¼ 2:44ðPðqÞTðqÞÞ2 Up
(9)
turn, would increase the efficiency of the hydrogen fuelled engine [4,15]. These are listed in Table 1.The efficiency and form factors for gasoline and hydrogen fuelling were found to be m ¼ 2 and a ¼ 5 [2,8,15].
In ref [13] the vanTyen Correlation was determined to be: 2 1 hðqÞ ¼ 5:41 104 PðqÞ3 TðqÞ3 3:19 þ 0:885Up
3.
Results and discussion
3.1.
Mass fraction burned
3.2.
Interpolated pressure model
(10)
Fig. 1 illustrates the comparative mass fraction burned models for both gasoline and hydrogen fuelled ICEs. For gasoline fuelling, it was found that MBT timing would produce maximum torque by employing a combustion duration of w70 that started at w30 BTDC and ended at w40 ATDC [6]. Unlike a gasoline fuelled ICE, a hydrogen fuelled ICE was found to have an MBT timing of w7 BTDC in order to produce maximum torque [15]. Therefore, the hydrogen fuelled ICE in this simulation was given a combustion duration of approxi mately 26 –28 starting at w8 BTDC and ending at approxi mately 18 –20 ATDC. The reason for retarding the timing to w8 BTDC for the hydrogen fuelled engine was due to the greater flame speed and for maximum torque [8]. The higher flame speed associated with the burning of hydrogen contributes to fast and turbulent combustion in the cylinder. That was found to contribute to the lowering of bsfc, which in
Fig. 2 illustrates the interpolated pressure model as developed and verified by Eriksson et al. and applied to both gasoline and hydrogen fuelled engines [5]. The lower cycle pressure associated with the hydrogen fuelled ICE is primarily due to the retarded spark timing and the decrease of moles associated with hydrogen combustion compared to the increase of moles associated with gasoline combustion. The effect of increasing moles is an amplification of the maximum pressure value calculated by the Ideal Gas Law. Conversely, the effect of the decreasing moles is a reduction of the maximum pressure value calculated by the Ideal Gas Law. Moreover, the effect of aggressive advancing or retarding of timing is a decrease in both cycle pressure and temperature [15]. It is known that hydrogen fuelled ICEs have lower cylinder pressures than their gasoline fuelled counterparts [7]. This decrease in MEP leads to a decrease in power output for the hydrogen fuelled engine for identical equivalence ratios. However, it seems to be possible to have a hydrogen fuelled engine operating at an equivalence ratio slightly greater than its gasoline counterpart to produce the same work output. Table 2 illustrates some pressures at critical points within the cycle such as: PIVC, PSOC, PMAx, and PEOC for an equivalence ratio of 1.0. Some
Table 3 – Comparative cycle temperatures for gasoline and hydrogen fuelling @ f [ 1.0. Crank Angle ( CA) Gasoline (K) Hydrogen (K)
qivc ¼ 240
qsoc
qcH2 ¼ 378 qcGAS ¼ 390
qeoc
qbdc ¼ 540
3320.180 3320.180
616.6275 726.4763
3326.3951 3488.8215
3182.2158 3488.8215
1714.2321 1537.6453
international journal of hydrogen energy 35 (2010) 5114–5123
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Fig. 4 – Gas speed (m/s) vs. crank angle (8CA) @ f [ 1.0 and 2500 rpms.
prescriptions to increase the power output of a hydrogen fuelled engine to that of its gasoline counterpart are: Increased compression ratios in order to take advantage of hydrogen’s ‘‘octane’’ number. Turbocharging/Supercharging in order to increase MEP and engine efficiency. Direct injection of H2 in order to avoid pre-ignition of fuel-air mixture due to increased working temperatures.
3.3.
Interpolated temperature model
Fig. 3 illustrates the comparative temperature model for both gasoline and hydrogen fuelled engines. This model behaves very similarly to the interpolated pressure model because it was constructed in a similar fashion on the premise that temperature and pressure are related via the Ideal Gas Law. It follows that if an accurate pressure model can be made; as in
ref [5], then an accurate temperature model can be made as well. The higher cylinder temperature associated with hydrogen combustion is primarily due to the higher LHV associated with hydrogen and the extended compression stroke [9]. However, the difference in maximum temperatures is less pronounced due to the lower specific fuel consumption coupled with the fuel-lean operation commonly associated with hydrogen fuelled ICEs [4,5]. The reason for the greater increase in temperature per unit crank angle associated with the hydrogen engine is due to the higher flame speed of hydrogen [9]. Table 3 lists several temperature values at key points throughout the cycle.
3.4.
Gas speed model
Fig. 4 illustrates the comparative in cylinder gas speeds for both gasoline and hydrogen fuelled engines for an equivalence ratio of 1.0 and an engine speed of 2500 rpms. A gas speed model was developed for both gasoline and hydrogen
Fig. 5 – Comparative P–V Diagram for gasoline and H2 fuelled engines @ f [ 1.0.
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Table 4a – Comparative Pressures and Volumes for Gasoline and H2 Fuelled 4-Stroke Otto Cycles @ f [ 0.4. Crank Angle ( CA) Gasoline Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
0 180 330 382 400 540 720
Volume (m3)
Pressure (kPa)
Hydrogen
0 180 352 378 378 540 720
Gasoline
Hydrogen
67.7266 67.7266 843.005 4344.230 2914.870 331.220 179.760
Gasoline
Hydrogen 5
67.7266 67.7266 1496.32 3303.77 3306.77 176.990 95.99
6.671 10 6.3371 104 1.0435 104 8.76112 105 1.3348 104 6.3371 104 6.6710 105
6.671 105 6.3371 104 6.937 105 8.0798 105 8.0798 105 6.3371 104 6.6710 105
Table 4b – Comparative Pressures and Volumes for Gasoline and H2 Fuelled 4-Stroke Otto Cycles @ f [ 1.0. Crank Angle ( CA)
Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
Volume (m3)
Pressure (kPa)
Gasoline
Hydrogen
Gasoline
Hydrogen
Gasoline
Hydrogen
0 180 330 382 400 540 720
0 180 352 378 378 540 720
67.7266 67.7266 843.005 7628.1027 5214.3716 590.2248 285.97
67.7266 67.7266 1496.32 6192.6083 6192.6083 334.0109 165.49
6.671 105 6.3371 104 1.0435 104 8.76112 105 1.3348 104 6.3371 104 6.6710 105
6.671 105 6.3371 104 6.937 105 8.0798 105 8.0798 105 6.3371 104 6.6710 105
Table 4c – Comparative Pressures and Volumes for Gasoline and H2 Fuelled 4-Stroke Otto Cycles @ f [ 1.2. Crank Angle ( CA) Gasoline Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
0 180 330 382 400 540 720
Hydrogen 0 180 352 378 378 540 720
Gasoline
Hydrogen
67.7266 67.7266 843.005 7794.23 5353.412 608.431 330.150
fuelling scenarios. In cylinder gas speed is heavily dependent on the pressure rise during combustion and the pressure drop during expansion. Furthermore, it was found to be constant as a function of mean piston speed during intake, exhaust and compression processes [2,3]. Since the cycle pressure of a gasoline fuelled ICE is greater than that of an equivalent hydrogen fuelled ICE therefore the in cylinder gas speed associated with a gasoline fuelled engine is greater than that of a hydrogen fuelled engine.
3.5.
Volume (m3)
Pressure (kPa)
Comparative P–V relationships
Fig. 5 illustrates the P–V diagrams for the 4-stroke Otto cycle for both gasoline and hydrogen fuelling scenarios equivalence ratios of 1.0. By integrating the P–V curve over the range of crank angles corresponding to the power stroke for both engines, it was found that the hydrogen fuelled engine should produce approximately 65% of the work output that a similar gasoline engine would normally produce at the same
Gasoline
Hydrogen 5
67.7266 67.7266 1496.32 6039.38 6039.38 340.63 179.160
6.671 10 6.3371 104 1.0435 104 8.76112 105 1.3348 104 6.3371 104 6.6710 105
6.671 105 6.3371 104 6.937 105 8.0798 105 8.079 105 6.3371 104 6.6710 105
equivalence ratio (f ¼ 1.0) [9]. Also, the hydrogen fuelled engine has a higher efficiency due to the heat addition being closer to TDC and under closer to isochoric conditions, resembling the Otto cycle [2,3,5]. Furthermore, due to the lesser exhaust blow down associated with the hydrogen fuelled engine the hydrogen engine should have a greater thermal efficiency [3]. Tables 4a–c list several pressure and volume values corresponding to both fuelling scenarios at equivalence ratios: f ¼ [0.4,1.0,1.2]. These equivalence ratios were chosen to understand the effects of fuel-lean, stoichiometric, and fuel-rich engine operation on the cycle pressure and by extension, work output.
3.6.
Comparative T–s relationships
Fig. 6 illustrates the T–s relationships for both gasoline and hydrogen fuelled engines. It can be seen that the heat addition for the hydrogen fuelled engine is greater than the heat addition for the gasoline fuelled engine, leading to higher
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Fig. 6 – Comparative T–s8 Diagrams for Gasoline and H2 Fuelled Engines @ f [ 1.0.
cylinder temperatures. Also, due to the higher combustion temperatures associated with the hydrogen fuelled engine there is a greater amount of entropy generation associated with that process [1]. Furthermore, the hydrogen fuelled engine exhibited a diesel like characteristic which accounts for a further increase in thermal efficiency. That is, it was found that the hydrogen fuelled engine rejected less heat to the environment on the exhaust stroke compared to the gasoline engine [2]. This translated into an increase in thermal efficiency by approximately 6.42% for the hydrogen fuelled ICE for an equivalence ratio of 1.0. Also, Tables 5a–c list comparative temperatures and entropies at 3 different
equivalence ratios of 0.4, 1.0, and 1.2 to simulate fuel-lean, stoichiometric, and fuel-rich combustion. The entropy values for gasoline combustion are in reasonable agreement with burned gas charts in ref [3] and therefore the extension of this method to calculate entropy values for hydrogen combustion should be within a similar degree of accuracy.
3.7.
Comparative heat transfer coefficients
In ref [13] by D’Souza et al., the convective heat transfer coefficients (HTCs) were calculated and graphed as a function of crank angle at 2500 rpm for both fuelling scenarios by
Table 5a – Comparative Temperatures and Entropies for Gasoline and H2 4-Stroke Otto Cycles @ f [ 0.4. Crank Angle ( CA) Gasoline Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
0 180 330 390 400 540 720
Temperature (K)
Hydrogen
0 180 352 378 378 540 720
Entropy (J/g*K)
Gasoline
Hydrogen
Gasoline
Hydrogen
298.01 300.0 616.628 1676.890 1614.043 855.884 344.134
297.77 300.0 726.476 1823.144 1823.144 801.933 323.293
6.943 6.948 6.992 8.561 8.500 8.341 7.489
7.900 7.910 7.938 8.961 8.961 8.774 7.927
Table 5b – Comparative Temperatures and Entropies for Gasoline and H2 4-Stroke Otto Cycles @ f [ 1.0. Crank Angle ( CA)
Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
Temperature (K)
Entropy (J/g*K)
Gasoline
Hydrogen
Gasoline
Hydrogen
Gasoline
Hydrogen
0 180 330 390 400 540 720
0 180 352 378 378 540 720
298.01 300.01 616.628 3326.3951 3182.2158 1714.2321 728.9530
297.77 300.01 726.476 3488.8215 3488.8215 1537.6453 653.8619
6.8507 6.8530 6.9358 9.4659 9.976 9.707 8.704
9.122 9.1344 9.1565 9.4626 9.4626 9.017 8.197
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Table 5c – Comparative temperatures and entropies for gasoline and H2 4-stroke otto cycles @ f [ 1.2. Crank Angle ( CA)
Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
Temperature (K)
Entropy (J/g*K)
Gasoline
Hydrogen
Gasoline
Hydrogen
Gasoline
Hydrogen
0 180 330 390 400 540 720
0 180 352 378 378 540 720
298.01 300.01 616.628 3566.60 3416.20 1842.59 743.50
297.77 300.01 726.476 3662.33 3662.33 1615.75 658.50
6.824 6.827 6.950 8.504 8.489 8.193 7.044
9.508 9.509 9.557 11.28 11.28 10.83 9.708
employing equations 5 through 10. These can be seen in Figs. 7 and 8 for gasoline and hydrogen fuelling scenarios, respectively. The heat transfer phenomenon in a spark ignition engine involves a very small amount of radiation of the gaseous and particulate matter to the cylinder walls in comparison to the convective heat transfer to the walls. Therefore, heat transfer due to radiation is considered negligible in this paper, as was mentioned in ref [2] as well. For a homogenous fuel-air mixture, the flame front propagates quickly across the combustion chamber. For example, the flame speed of hydrogen was found to be 2.1 m/s which is greater than the flame speed of 0.35 m/s associated with gasoline [9]. The radiant heat transfer from this flame does reduce in cylinder temperature. To incorporate this in the
calculations of heat transfer for a hydrogen and gasoline engine, the Woschni correlation is used. The Woschni correlation is based on heat flux and therefore, the radiation heat transfer to the cylinder wall is accounted for. Other correlations, like the Annand correlation does not include the radiation heat flux and hence the radiant energy absorbed, emitted, etc with respect to solid angle direction must be included in the final heat transfer equation. Furthermore, it is fair to mention the inapplicability of the Eichelberg correlation in this study. Since the Eichelberg correlation was based on a solution to a temperature variation of the heat transfer rates
Fig. 7 – a. Select comparative HTCs for gasoline engine operating at 2500 rpm [13], b. All comparative HTCs for gasoline engine operating at 2500 rpms [13].
Fig. 8 – a. Select Comparative HTCs for H2 Engine operating at 2500 rpms [13], b. All Comparative HTCs for H2 Engine operating at 2500 rpms [13].
international journal of hydrogen energy 35 (2010) 5114–5123
from a 2 stroke diesel engine, by definition it could not be expected to model the heat transfer coefficient of a 4-stroke gasoline or hydrogen engine accurately. The results show that the Eichelberg correlation undershoots the estimation of convective heat transfer coefficient modelled by Woschni’s correlation and all other correlations.
4.
Conclusions
The analysis of the comparative combustion characteristics of gasoline and hydrogen fuelled internal combustion engines was able to elucidate on the potential performance and efficiency of a hydrogen fuelled ICE compared to a gasoline fuelled ICE. It was found that a hydrogen fuelled ICE had a higher thermal efficiency by approximately 6.42% due to less heat rejection during the exhaust stroke, less blow down during the exhaust stroke, combustion taking place closer to TDC and combustion taking place in an closer to isochoric environment and thus, closer to an actual Otto cycle. Finally, the hydrogen fuelled ICE was found to have 65% of the work output regularly associated with a gasoline fuelled ICE due to decreased MEP and increased heat transfer to the cooling system. The greater amount of convective heat transfer in the hydrogen fuelled ICE is postulated to occur due to the higher flame speed which causes faster flame propagation and therefore a more turbulent combustion reaction. Also, the higher combustion temperatures and shorter quenching distance associated with hydrogen combustion are believed to be cause to greater convective heat transfer to the cylinder walls.
Acknowledgements The support for this work provided by the Natural Sciences and Engineering Research Council of Canada is gratefully acknowledged.
Nomenclature Abbreviations and symbols A Exposed area (m2) ATDC After top dead Center b Bore (m) BDC Bottom dead center BTDC Before top dead center bsfc Brake specific fuel consumption EOC End of combustion exh Exhaust exp Expansion h Convective heat transfer coefficient (W/m2K) HTC Heat transfer coefficient (W/m2K) ICE Internal combustion engine k Ratio of specific heats (1.40) LHV Lower heating value MEP Mean effective pressure (kPa) P Pressure (kPa) R Ratio of Connecting Rod Length to Stroke Residual gas fraction (-) fr s Stroke (m) Specific entropy (J/g*K) s
SOC T TDC U Up V Vc Vd
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Start of combustion Temperature ( K) Top dead center Gas speed (cm/s) Mean piston speed (m/s) Volume (m3) Clearance volume (m3) Displacement volume (m3)
Subscripts and superscripts 1 State Point 1 of otto cycle (intake) 2 State point 2 of otto cycle (compression) 3 State point 3 of otto cycle (maximum value) 4 State point 4 of otto cycle (exhaust stroke) Greek symbols q Degrees of crank angle ( C.A) Dq Combustion duration ( C.A) f Equivalence ratio (-)
references
[1] Cengel Yunus, Boles Micheal. Thermodynamics: an engineering approach. McGraw-Hill Education; 2005. [2] Ferguson Colin, R, Kirkpatrick Allan, T. Internal combustion engines: applied thermosciences. Wiley; 2004. [3] Heywood John. Internal combustion engine fundamentals. McGraw-Hill Education; 1988. [4] Galloni, Enzo. Combustion Modelling and Performance Estimation of an S.I Engine Using Lean Hydrogen-Air Mixtures, WHEC 16 In: Proceedings, p. 1–6; June 2006. [5] Eriksson Lars, Andersson Ingemar. An analytic model for cylinder pressure in a four stroke S.I engine. SAE; 2002. 200201-0371, pp.1–9. [6] Zhang, S; Sobiesiak, A. The first and second law analyses of a port injected, spark ignition engine fuelled with compressed natural gas (CNG); 2002. [7] Tomita Eiji. Hydrogen internal combustion engines. Academic report. Okayama: Okayama University; 2001. [8] Szwaja S, Bhandary KR, Naber JD. Comparisons of hydrogen and gasoline combustion knock in a spark ignition engine. International Association for Hydrogen Energy 2007: 1-. [9] College of Desert. Module 3: hydrogen use in internal combustion engines, http://www.eere.energy.gove/ hydrogenandfuelcells/tech_validation/pdfs/fcm03r0.pdf; December 2001 [accessed 25.09.08]. [10] Nijeweme DJ, Oude JBW, Kok CR, StoneWyszynski L. Unsteady in-cylinder heat transfer in a spark ignition engine: experiments and modelling. Pros. Instn. Mechanical Engineers 2001;215(Part D):747–60. [11] Suzuki Hiroyuki, Shudo Toshio. Applicability of heat transfer equations to hydrogen combustion. JSAE Review 2002;23(3): 303–8. [12] Sharief A, Chandrashekar TK, Antony AJ, Samaga BS. Study on heat transfer correlation in IC engines. SAE International; 2008. 1–8. [13] D’Souza, Ninochka. Comparative heat transfer analysis of gasoline and hydrogen fuelled ICEs. UOIT B.Eng Thesis Report; 2009. [14] Nieminen, Jonathan. Comparative Exergy Analyses of Gasoline and Hydrogen Fuelled ICEs. ICH2P-09 In: Proceedings; May 2009. [15] Halmari, Jaakko. Computer Simulations of a Hydrogen Fuelled Internal Combustion Engine. Texas Technical University Thesis Report; May 2005.
Available at www.sciencedirect.com
journal homepage: www.elsevier.com/locate/he
Comparative combustion characteristics of gasoline and hydrogen fuelled ICEs Jonathan Nieminen*, Ninochka D’Souza, Ibrahim Dincer Faculty of Engineering and Applied Science, University of Ontario Institute of Technology, 2000 Simcoe Street North, Oshawa, ON L1H 7K4, Canada
article info
abstract
Article history:
In this paper, some models of comparative combustion characteristics for gasoline and
Received 14 July 2009
hydrogen fuelled spark ignition internal combustion engines were developed and dis-
Received in revised form
cussed from a thermodynamic and heat transfer perspective. The geometry used was that
15 August 2009
of a 3.4L GM V6 engine with a compression ratio of 9.5:1. Models for mass fraction burned,
Accepted 30 August 2009
pressure, temperature, and gas speed were developed according to the literature survey
Available online 2 October 2009
and graphed over the cycle range. Furthermore, Pressure–Volume and Temperature– Entropy models were developed for both gasoline and hydrogen fuelled engines. Analysis
Keywords:
of these models indicated approximately a 6.42% increase thermal efficiency for the
Hydrogen
hydrogen fuelled engine due to less exhaust blow down, less heat rejection during the
Gasoline
exhaust stroke, and its shorter combustion duration closer to TDC. However, it was found
Engines
that the hydrogen fuelled engine had approximately a 35.0% decrease in power output at
Thermodynamic
an equivalence ratio of 1.0 due to the decrease in MEP and a greater amount of heat transfer
Heat Transfer
to the cooling system due to the increased combustion temperatures, shorter quenching distance associated with H2 combustion and greater flame speed. Finally, an increase in cycle temperatures and pressures was observed from increasing the equivalence ratio from 0.4 to 1.0 to 1.2. ª 2009 Professor T. Nejat Veziroglu. Published by Elsevier Ltd. All rights reserved.
1.
Introduction
Due to rising fuel costs, stricter environmental regulations, decreasing oil supplies, and additional problems facing modern society, alternative fuels such as hydrogen have gained popularity. However, that alone is insufficient reason to develop such a technology. The development of a hydrogen internal combustion engine requires a solid understanding of the comparative performance characteristics of the engine. Therefore, comparative pressure models for both gasoline and hydrogen fuelled ICEs are developed as in ‘‘An Analytic Model for Cylinder Pressure in a Four Stroke S.I Engine’’. Since the
pressure and temperature of a gas are related via the Ideal Gas Law; therefore the extension of the pressure model to develop a comparative temperature model for both gasoline and hydrogen fuelling is considered to be justified and accurate. The pressure model will be graphed against volume model in order to make a Pressure–Volume model for both gasoline and hydrogen fuelling. This comparative model will then be used to determine the effect of differing cycle pressure on work output and thermal efficiency between a gasoline and hydrogen fuelled engine. Similarly, the temperature model will be graphed against the entropy generation model in order to make a Temperature–Entropy model for both gasoline and
* Corresponding author. E-mail addresses: [email protected] (J. Nieminen), [email protected] (N. D’Souza), [email protected] (I. Dincer). 0360-3199/$ – see front matter ª 2009 Professor T. Nejat Veziroglu. Published by Elsevier Ltd. All rights reserved. doi:10.1016/j.ijhydene.2009.08.098
international journal of hydrogen energy 35 (2010) 5114–5123
hydrogen fuelling. This comparative model will be used to determine the effect of cycle temperature on thermal efficiency. Also, models for cylinder volume, exposed wall area, and gas speed as a function of crank angle were found [2]. Plentiful background information relating to the thermodynamic analysis of spark ignition internal combustion engines is available and has been used to aid this analysis. In ‘‘An Analytic Model for Cylinder Pressure in a Four Stroke S.I Engine’’ by Eriksson et al. effectively developed a model for cylinder pressure for by interpolating between compression and expansion asymptotes as defined by the Otto Cycle. The reason behind such an interpolation is the result of the heat addition not being isochoric or isobaric. Therefore a model was developed to compensate by the incorporation of the Wiebe Burn Function over the combustion duration in order to interpolate between the polytropic compression and expansion asymptotes. Varying cylinder pressure during combustion is modelled with varying volume as a function of burning rate. The compression asymptote was modelled as a polytropic compression from qbdc to qsoc, and the expansion asymptote was modelled as a polytropic expansion from qeoc to qbdc. The range between qsoc and qeoc is described accurately through the interpolated pressure function. According to Enzo Galloni in ‘‘Combustion Modelling and Performance Estimation of an S.I Engine Using Lean H2-Air Mixtures’’ developed a numerical model that predicted that an ultra-lean, fast burning H2 could be able to control engine load by varying the amount of fuel at wide open throttle (WOT) conditions, thus limiting engine pumping losses. Galloni [4] eluded that this decrease in pumping losses would translate into an increased thermal efficiency. All of this was hypothesized by the model then proven to through experimentation. When compared to gasoline fuelling there are higher cylinder pressures for H2 fuelling exist during unthrottled operation. These higher cylinder pressures were mentioned as a potential safety hazard to due increased mechanical and thermal stresses. ‘‘Hydrogen Internal Combustion Engines’’ by Eiji Tomita covers the principles of H2 fuelling for internal combustion engines. Pressure and knock profiles were constructed for various fuels such as gasoline, n-butane, and hydrogen. It was found that gasoline had the highest cylinder pressures and hydrogen had the lowest. Also, hydrogen and gasoline were found to have similar knock profiles that differ by timing. The hydrogen knock profile was found to reach its maximum value later in the cycle whereas the gasoline knock profile was found to reach its maximum value earlier in the cycle. The latter is more damaging to the engine because it is taking place at higher pressures. According to Szwaja et al. in ‘‘Comparison of Hydrogen and Gasoline Combustion Knock in an Internal Combustion Engine’’ the ignition timing of a hydrogen engine was advanced to approximately 5 to 10 BTDC. Also, they discussed the effects of combustion timing on knock in both hydrogen and gasoline fuelled internal combustion engines. Along with other literature, they had a spark timing of approximately 5 to 10 ATDC and had been found to end around 18 to 20 ATDC. Due to the greater flame speed of hydrogen the spark timing is moved closer to TDC. This yields a combustion duration of approximately 26 to 28 . Finally, a full computer simulation of a hydrogen fuelled, spark ignition ICE was done in ref [15]. Halmari [15] showed the effect of
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retarded and advanced timing on a hydrogen fuelled ICE, simulated engine torque as a function of spark timing, and brake thermal efficiency as a function of equivalence ratio for a supercharged at various compression ratios. It was found that brake torque was a maximum at around 7 BTDC, the brake thermal efficiency peaked at a compression ratio of approximately 12.6:1 and an equivalence ratio of approximately 0.9. With respect to heat transfer within internal combustion engines, much work has been done in the area and was helpful in understanding the phenomenon. Both Kirkpatrick et al. and Heywood in their respective books devote chapters towards the understanding of heat transfer in internal combustion engines, ref. [2] and ref. [3] respectively. According to Shudo et al. in ‘‘Applicability of Heat Transfer Equations to Hydrogen Combustion’’, hydrogen combustion exhibits greater heat loss to the coolant due to higher flame speed and shorter quenching distance. According to Nijeweme et al. in ‘‘Unsteady In Cylinder Heat Transfer in a Spark Ignition Engine: Experiments and Modelling’’ [10], evaluating these temperatures, fluxes, limits, etc not only helps in improving the efficiency of engine performance but it also has a huge impact on exhaust emissions that are exponentially dependant on the peak combustion temperature values. In this paper, comparative combustion characteristics of gasoline and hydrogen fuelled internal combustion engines (ICEs) such as – mass fraction burned (mfb), pressure, temperature, and gas speed. Furthermore, Pressure–Volume and Temperature–Entropy diagrams were developed and commented upon from a thermodynamic and heat transfer perspective.
1.1.
Thermodynamic analysis
To construct a thermodynamic model for the working fluid inside an internal combustion engine, following assumptions are made:
Spark ignition engine operating on the 4-stroke Otto Cycle. Working fluids are ideal gases. The compression and expansion processes are polytropic. Neglecting the effects of friction, valve and piston ring blowby, squish and swirl.
1.2.
Mass fraction burned
The mass fraction burned (mfb) represents the percentage of fuel burned as a function of crank angle throughout the combustion duration. It is modelled by a Weibe Burn Function with m ¼ 2 and a ¼ 5 [2]. q qSOC ; qSOC < q < qEOC f ðqÞ ¼ 1 exp a Dq
1.3.
(1)
Interpolated pressure model
Since combustion in a spark ignition engine takes place over a period of time known as combustion duration it is not isochoric as the Otto cycle suggests. Therefore, development of an accurate model of the behaviour of the pressure during the
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international journal of hydrogen energy 35 (2010) 5114–5123
Fig. 1 – Mass fraction burned (%/100) vs. crank angle (8CA).
combustion process is fundamental to an accurate model of the cycle as a whole [5]. 8 > > > > > > > > > > <
P1
k
vd P1 vðqÞ k k vd vc PðqÞ ¼ ð1 f ðqÞÞP1 vðqÞ þf ðqÞP3 vðqÞ > k > > vc > P3 vðqÞ > > > > k > > vd : P4 vðqÞ
; qTDC q < qBDC ; qBDC q < qSOC ; qSOC q < qEOC
(2)
; qEOC q < qBDC
1.5.
Gas speed in an ICE varies changes with pressure and piston speed. Gas speed is heavily dependent on the pressure rise during combustion. It was modelled as a piecewise function based on whether or not the valves are open or closed and which in cylinder processes are taking place. The in cylinder gas speed can be modelled according to refs [2,3].
; qBDC q < qTDC
1.6. 1.4.
Gas speed
Interpolated temperature model
Since heat release in a real SI engine is not isochoric as the Otto cycle suggests and since temperature and pressure are related via the Ideal Gas Law, therefore the behaviour of the interpolated temperature model is accurate by being a logical extension of the interpolated pressure model. The temperature during the intake stroke was modelled as in refs [2,3] in order to account for the residual fraction of combustion products still in the cylinder after the exhaust stroke. The residual fractions of burned gases were found to be 0.01166 (1.166%) and 0.011826 (1.1826%) for hydrogen and gasoline, respectively. 8 k1 i h k p1 > ; qTDC q < qBDC 1 fr T1 þ fr Texh ðqÞ 1 1 pexh > > ðqÞ > > k1 > > vd > > T1 vðqÞ ; qBDC q < qSOC > > < k1 k1 v v c TðqÞ ¼ ð1 f ðqÞÞT1 d þf ðqÞT3 vðqÞ ; qSOC q < qEOC vðqÞ > > k1 > > > vc > T3 vðqÞ ; qEOC q < qBDC > > > > k1 > : vðqÞ T4 vd ; qBDC q < qTDC
Comparative pressure–volume relationships
A Pressure–Volume (P–V) model is essential to measuring the work output of a thermodynamic gas power cycle. Since gasoline and hydrogen fuelled engines have differing pressure relationships, it follows that the work output of the cycle would be affected. The work done by pressure over the cycle is called the mean effective pressure (MEP). In general, a greater MEP is indicative of a greater work output of the engine. The interpolated pressure model was extended to accommodate exhaust and intake processes. There are 4 processes in a 4stroke Otto cycle [2,3]:
Intake Stroke from qTDC to qBDC Polytropic Compression from qBDC to qTDC Polytropic Expansion from qTDC to qBDC Exhaust Stroke from qBDC to qTDC
1.7.
Comparative temperature – entropy relationships
A Temperature–Entropy (T–s ) model is essential in understanding the amount of heat added or lost during a process and/
(3) where:
fr ¼
1 pexh rc pexp
Table 1 – Combustion durations and ignition timings. !1k (4)
Fuel
qsoc
qeoc
6q
Hydrogen Gasoline
352 330
378 400
26 70
international journal of hydrogen energy 35 (2010) 5114–5123
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Fig. 2 – Cylinder pressure (kPa) vs. crank angle (8CA).
or how irreversible a thermodynamic process is by its deviation from the isentropic lines. The interpolated temperature model was employed here. Furthermore, entropy was found to be a function of the pressure, temperature, and equivalence ratio of the mixture over the entire cycle range [3,14].
law of similarity [12] and proposed a function based on pressure, cylinder bore, gas speed, and temperature [2]. When applied to gasoline ICEs as a function of crank angle it was found to be: hðqÞ ¼ 3:26PðqÞ0:8 UðqÞ0:8 b0:2 TðqÞ0:55
2.
Heat transfer review
Heat transfer within an ICE is directly related to its thermal performance. For the goal and scope of this light analysis, it is reasonable to make a few assumptions. Convection is the dominant mode of heat transfer within the cylinder[2]. Simplistic ‘‘single zone’’ combustion model. Finite heat release method is assumed to be valid.
However, when the aforementioned formula is applied to hydrogen fuelled ICEs it undershoots the amount of heat transfer to the coolant by approximately 75% [11]. To offset this, the convective heat transfer coefficient in Woschni’s correlation was multiplied by 4 to estimate the actual amount of convective heat transfer more accurately. When made a function of crank angle, the Woschni’s correlation took the form of: hðqÞ ¼ 13:04PðqÞ0:8 UðqÞ0:8 b0:2 TðqÞ0:55
2.3. 2.1.
(6)
Other correlations
Comparative heat transfer coefficients
The in cylinder processes change with respect to the crank angle and the nature of the process occurring within the 4stroke cycle. For internal combustion engines, there are several relations for convective heat transfer coefficients that have been calculated through experimentation and/or analysis.
2.2.
(5)
The Nusselt Correlation was constructed from the data on quiescent heat loss measurements from a spherical bomb combustor [13]. It took the form of: 1=3 (7) hðqÞ ¼ 1:151 1 þ 1:24Up PðqÞ2 TðqÞ The Briling Correlation was derived as a slight modification of the mean speed factor term in the Nusselt Correlation [13]. 1=3 hðqÞ ¼ 1:151 3:5 þ 0:185Up PðqÞ2 TðqÞ
Woschni’s correlation
Woschni reviewed various heat transfer correlations for reciprocating engines and proposed a formula based on the
(8)
The Eichelberg Correlation was determined through the measurements of instantaneous heat transfer rates from a 2
Table 2 – Comparative cycle pressures for gasoline and H2 fuelling @ f [ 1.0. Crank Angle ( CA) Gasoline (kPa) Hydrogen (kPa)
qivc ¼ 240
qsoc
qcH2 ¼ 378 qcGAS ¼ 382
qeoc
qbdc ¼ 540
96.451 96.451
843.005 1496.32
7628.1027 6192.6083
5214.3716 6192.6083
590.2246 334.0109
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international journal of hydrogen energy 35 (2010) 5114–5123
Fig. 3 – Cylinder temperature (K) vs. crank angle (8CA) @ f [ 1.0.
stroke diesel engine [13]. Through solving the temperature variation through a harmonic analysis, he concluded that: 1=3 1 hðqÞ ¼ 2:44ðPðqÞTðqÞÞ2 Up
(9)
turn, would increase the efficiency of the hydrogen fuelled engine [4,15]. These are listed in Table 1.The efficiency and form factors for gasoline and hydrogen fuelling were found to be m ¼ 2 and a ¼ 5 [2,8,15].
In ref [13] the vanTyen Correlation was determined to be: 2 1 hðqÞ ¼ 5:41 104 PðqÞ3 TðqÞ3 3:19 þ 0:885Up
3.
Results and discussion
3.1.
Mass fraction burned
3.2.
Interpolated pressure model
(10)
Fig. 1 illustrates the comparative mass fraction burned models for both gasoline and hydrogen fuelled ICEs. For gasoline fuelling, it was found that MBT timing would produce maximum torque by employing a combustion duration of w70 that started at w30 BTDC and ended at w40 ATDC [6]. Unlike a gasoline fuelled ICE, a hydrogen fuelled ICE was found to have an MBT timing of w7 BTDC in order to produce maximum torque [15]. Therefore, the hydrogen fuelled ICE in this simulation was given a combustion duration of approxi mately 26 –28 starting at w8 BTDC and ending at approxi mately 18 –20 ATDC. The reason for retarding the timing to w8 BTDC for the hydrogen fuelled engine was due to the greater flame speed and for maximum torque [8]. The higher flame speed associated with the burning of hydrogen contributes to fast and turbulent combustion in the cylinder. That was found to contribute to the lowering of bsfc, which in
Fig. 2 illustrates the interpolated pressure model as developed and verified by Eriksson et al. and applied to both gasoline and hydrogen fuelled engines [5]. The lower cycle pressure associated with the hydrogen fuelled ICE is primarily due to the retarded spark timing and the decrease of moles associated with hydrogen combustion compared to the increase of moles associated with gasoline combustion. The effect of increasing moles is an amplification of the maximum pressure value calculated by the Ideal Gas Law. Conversely, the effect of the decreasing moles is a reduction of the maximum pressure value calculated by the Ideal Gas Law. Moreover, the effect of aggressive advancing or retarding of timing is a decrease in both cycle pressure and temperature [15]. It is known that hydrogen fuelled ICEs have lower cylinder pressures than their gasoline fuelled counterparts [7]. This decrease in MEP leads to a decrease in power output for the hydrogen fuelled engine for identical equivalence ratios. However, it seems to be possible to have a hydrogen fuelled engine operating at an equivalence ratio slightly greater than its gasoline counterpart to produce the same work output. Table 2 illustrates some pressures at critical points within the cycle such as: PIVC, PSOC, PMAx, and PEOC for an equivalence ratio of 1.0. Some
Table 3 – Comparative cycle temperatures for gasoline and hydrogen fuelling @ f [ 1.0. Crank Angle ( CA) Gasoline (K) Hydrogen (K)
qivc ¼ 240
qsoc
qcH2 ¼ 378 qcGAS ¼ 390
qeoc
qbdc ¼ 540
3320.180 3320.180
616.6275 726.4763
3326.3951 3488.8215
3182.2158 3488.8215
1714.2321 1537.6453
international journal of hydrogen energy 35 (2010) 5114–5123
5119
Fig. 4 – Gas speed (m/s) vs. crank angle (8CA) @ f [ 1.0 and 2500 rpms.
prescriptions to increase the power output of a hydrogen fuelled engine to that of its gasoline counterpart are: Increased compression ratios in order to take advantage of hydrogen’s ‘‘octane’’ number. Turbocharging/Supercharging in order to increase MEP and engine efficiency. Direct injection of H2 in order to avoid pre-ignition of fuel-air mixture due to increased working temperatures.
3.3.
Interpolated temperature model
Fig. 3 illustrates the comparative temperature model for both gasoline and hydrogen fuelled engines. This model behaves very similarly to the interpolated pressure model because it was constructed in a similar fashion on the premise that temperature and pressure are related via the Ideal Gas Law. It follows that if an accurate pressure model can be made; as in
ref [5], then an accurate temperature model can be made as well. The higher cylinder temperature associated with hydrogen combustion is primarily due to the higher LHV associated with hydrogen and the extended compression stroke [9]. However, the difference in maximum temperatures is less pronounced due to the lower specific fuel consumption coupled with the fuel-lean operation commonly associated with hydrogen fuelled ICEs [4,5]. The reason for the greater increase in temperature per unit crank angle associated with the hydrogen engine is due to the higher flame speed of hydrogen [9]. Table 3 lists several temperature values at key points throughout the cycle.
3.4.
Gas speed model
Fig. 4 illustrates the comparative in cylinder gas speeds for both gasoline and hydrogen fuelled engines for an equivalence ratio of 1.0 and an engine speed of 2500 rpms. A gas speed model was developed for both gasoline and hydrogen
Fig. 5 – Comparative P–V Diagram for gasoline and H2 fuelled engines @ f [ 1.0.
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Table 4a – Comparative Pressures and Volumes for Gasoline and H2 Fuelled 4-Stroke Otto Cycles @ f [ 0.4. Crank Angle ( CA) Gasoline Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
0 180 330 382 400 540 720
Volume (m3)
Pressure (kPa)
Hydrogen
0 180 352 378 378 540 720
Gasoline
Hydrogen
67.7266 67.7266 843.005 4344.230 2914.870 331.220 179.760
Gasoline
Hydrogen 5
67.7266 67.7266 1496.32 3303.77 3306.77 176.990 95.99
6.671 10 6.3371 104 1.0435 104 8.76112 105 1.3348 104 6.3371 104 6.6710 105
6.671 105 6.3371 104 6.937 105 8.0798 105 8.0798 105 6.3371 104 6.6710 105
Table 4b – Comparative Pressures and Volumes for Gasoline and H2 Fuelled 4-Stroke Otto Cycles @ f [ 1.0. Crank Angle ( CA)
Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
Volume (m3)
Pressure (kPa)
Gasoline
Hydrogen
Gasoline
Hydrogen
Gasoline
Hydrogen
0 180 330 382 400 540 720
0 180 352 378 378 540 720
67.7266 67.7266 843.005 7628.1027 5214.3716 590.2248 285.97
67.7266 67.7266 1496.32 6192.6083 6192.6083 334.0109 165.49
6.671 105 6.3371 104 1.0435 104 8.76112 105 1.3348 104 6.3371 104 6.6710 105
6.671 105 6.3371 104 6.937 105 8.0798 105 8.0798 105 6.3371 104 6.6710 105
Table 4c – Comparative Pressures and Volumes for Gasoline and H2 Fuelled 4-Stroke Otto Cycles @ f [ 1.2. Crank Angle ( CA) Gasoline Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
0 180 330 382 400 540 720
Hydrogen 0 180 352 378 378 540 720
Gasoline
Hydrogen
67.7266 67.7266 843.005 7794.23 5353.412 608.431 330.150
fuelling scenarios. In cylinder gas speed is heavily dependent on the pressure rise during combustion and the pressure drop during expansion. Furthermore, it was found to be constant as a function of mean piston speed during intake, exhaust and compression processes [2,3]. Since the cycle pressure of a gasoline fuelled ICE is greater than that of an equivalent hydrogen fuelled ICE therefore the in cylinder gas speed associated with a gasoline fuelled engine is greater than that of a hydrogen fuelled engine.
3.5.
Volume (m3)
Pressure (kPa)
Comparative P–V relationships
Fig. 5 illustrates the P–V diagrams for the 4-stroke Otto cycle for both gasoline and hydrogen fuelling scenarios equivalence ratios of 1.0. By integrating the P–V curve over the range of crank angles corresponding to the power stroke for both engines, it was found that the hydrogen fuelled engine should produce approximately 65% of the work output that a similar gasoline engine would normally produce at the same
Gasoline
Hydrogen 5
67.7266 67.7266 1496.32 6039.38 6039.38 340.63 179.160
6.671 10 6.3371 104 1.0435 104 8.76112 105 1.3348 104 6.3371 104 6.6710 105
6.671 105 6.3371 104 6.937 105 8.0798 105 8.079 105 6.3371 104 6.6710 105
equivalence ratio (f ¼ 1.0) [9]. Also, the hydrogen fuelled engine has a higher efficiency due to the heat addition being closer to TDC and under closer to isochoric conditions, resembling the Otto cycle [2,3,5]. Furthermore, due to the lesser exhaust blow down associated with the hydrogen fuelled engine the hydrogen engine should have a greater thermal efficiency [3]. Tables 4a–c list several pressure and volume values corresponding to both fuelling scenarios at equivalence ratios: f ¼ [0.4,1.0,1.2]. These equivalence ratios were chosen to understand the effects of fuel-lean, stoichiometric, and fuel-rich engine operation on the cycle pressure and by extension, work output.
3.6.
Comparative T–s relationships
Fig. 6 illustrates the T–s relationships for both gasoline and hydrogen fuelled engines. It can be seen that the heat addition for the hydrogen fuelled engine is greater than the heat addition for the gasoline fuelled engine, leading to higher
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international journal of hydrogen energy 35 (2010) 5114–5123
Fig. 6 – Comparative T–s8 Diagrams for Gasoline and H2 Fuelled Engines @ f [ 1.0.
cylinder temperatures. Also, due to the higher combustion temperatures associated with the hydrogen fuelled engine there is a greater amount of entropy generation associated with that process [1]. Furthermore, the hydrogen fuelled engine exhibited a diesel like characteristic which accounts for a further increase in thermal efficiency. That is, it was found that the hydrogen fuelled engine rejected less heat to the environment on the exhaust stroke compared to the gasoline engine [2]. This translated into an increase in thermal efficiency by approximately 6.42% for the hydrogen fuelled ICE for an equivalence ratio of 1.0. Also, Tables 5a–c list comparative temperatures and entropies at 3 different
equivalence ratios of 0.4, 1.0, and 1.2 to simulate fuel-lean, stoichiometric, and fuel-rich combustion. The entropy values for gasoline combustion are in reasonable agreement with burned gas charts in ref [3] and therefore the extension of this method to calculate entropy values for hydrogen combustion should be within a similar degree of accuracy.
3.7.
Comparative heat transfer coefficients
In ref [13] by D’Souza et al., the convective heat transfer coefficients (HTCs) were calculated and graphed as a function of crank angle at 2500 rpm for both fuelling scenarios by
Table 5a – Comparative Temperatures and Entropies for Gasoline and H2 4-Stroke Otto Cycles @ f [ 0.4. Crank Angle ( CA) Gasoline Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
0 180 330 390 400 540 720
Temperature (K)
Hydrogen
0 180 352 378 378 540 720
Entropy (J/g*K)
Gasoline
Hydrogen
Gasoline
Hydrogen
298.01 300.0 616.628 1676.890 1614.043 855.884 344.134
297.77 300.0 726.476 1823.144 1823.144 801.933 323.293
6.943 6.948 6.992 8.561 8.500 8.341 7.489
7.900 7.910 7.938 8.961 8.961 8.774 7.927
Table 5b – Comparative Temperatures and Entropies for Gasoline and H2 4-Stroke Otto Cycles @ f [ 1.0. Crank Angle ( CA)
Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
Temperature (K)
Entropy (J/g*K)
Gasoline
Hydrogen
Gasoline
Hydrogen
Gasoline
Hydrogen
0 180 330 390 400 540 720
0 180 352 378 378 540 720
298.01 300.01 616.628 3326.3951 3182.2158 1714.2321 728.9530
297.77 300.01 726.476 3488.8215 3488.8215 1537.6453 653.8619
6.8507 6.8530 6.9358 9.4659 9.976 9.707 8.704
9.122 9.1344 9.1565 9.4626 9.4626 9.017 8.197
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international journal of hydrogen energy 35 (2010) 5114–5123
Table 5c – Comparative temperatures and entropies for gasoline and H2 4-stroke otto cycles @ f [ 1.2. Crank Angle ( CA)
Qin Qcomp Qsoc Qmax Qeoc Qexp Qexh
Temperature (K)
Entropy (J/g*K)
Gasoline
Hydrogen
Gasoline
Hydrogen
Gasoline
Hydrogen
0 180 330 390 400 540 720
0 180 352 378 378 540 720
298.01 300.01 616.628 3566.60 3416.20 1842.59 743.50
297.77 300.01 726.476 3662.33 3662.33 1615.75 658.50
6.824 6.827 6.950 8.504 8.489 8.193 7.044
9.508 9.509 9.557 11.28 11.28 10.83 9.708
employing equations 5 through 10. These can be seen in Figs. 7 and 8 for gasoline and hydrogen fuelling scenarios, respectively. The heat transfer phenomenon in a spark ignition engine involves a very small amount of radiation of the gaseous and particulate matter to the cylinder walls in comparison to the convective heat transfer to the walls. Therefore, heat transfer due to radiation is considered negligible in this paper, as was mentioned in ref [2] as well. For a homogenous fuel-air mixture, the flame front propagates quickly across the combustion chamber. For example, the flame speed of hydrogen was found to be 2.1 m/s which is greater than the flame speed of 0.35 m/s associated with gasoline [9]. The radiant heat transfer from this flame does reduce in cylinder temperature. To incorporate this in the
calculations of heat transfer for a hydrogen and gasoline engine, the Woschni correlation is used. The Woschni correlation is based on heat flux and therefore, the radiation heat transfer to the cylinder wall is accounted for. Other correlations, like the Annand correlation does not include the radiation heat flux and hence the radiant energy absorbed, emitted, etc with respect to solid angle direction must be included in the final heat transfer equation. Furthermore, it is fair to mention the inapplicability of the Eichelberg correlation in this study. Since the Eichelberg correlation was based on a solution to a temperature variation of the heat transfer rates
Fig. 7 – a. Select comparative HTCs for gasoline engine operating at 2500 rpm [13], b. All comparative HTCs for gasoline engine operating at 2500 rpms [13].
Fig. 8 – a. Select Comparative HTCs for H2 Engine operating at 2500 rpms [13], b. All Comparative HTCs for H2 Engine operating at 2500 rpms [13].
international journal of hydrogen energy 35 (2010) 5114–5123
from a 2 stroke diesel engine, by definition it could not be expected to model the heat transfer coefficient of a 4-stroke gasoline or hydrogen engine accurately. The results show that the Eichelberg correlation undershoots the estimation of convective heat transfer coefficient modelled by Woschni’s correlation and all other correlations.
4.
Conclusions
The analysis of the comparative combustion characteristics of gasoline and hydrogen fuelled internal combustion engines was able to elucidate on the potential performance and efficiency of a hydrogen fuelled ICE compared to a gasoline fuelled ICE. It was found that a hydrogen fuelled ICE had a higher thermal efficiency by approximately 6.42% due to less heat rejection during the exhaust stroke, less blow down during the exhaust stroke, combustion taking place closer to TDC and combustion taking place in an closer to isochoric environment and thus, closer to an actual Otto cycle. Finally, the hydrogen fuelled ICE was found to have 65% of the work output regularly associated with a gasoline fuelled ICE due to decreased MEP and increased heat transfer to the cooling system. The greater amount of convective heat transfer in the hydrogen fuelled ICE is postulated to occur due to the higher flame speed which causes faster flame propagation and therefore a more turbulent combustion reaction. Also, the higher combustion temperatures and shorter quenching distance associated with hydrogen combustion are believed to be cause to greater convective heat transfer to the cylinder walls.
Acknowledgements The support for this work provided by the Natural Sciences and Engineering Research Council of Canada is gratefully acknowledged.
Nomenclature Abbreviations and symbols A Exposed area (m2) ATDC After top dead Center b Bore (m) BDC Bottom dead center BTDC Before top dead center bsfc Brake specific fuel consumption EOC End of combustion exh Exhaust exp Expansion h Convective heat transfer coefficient (W/m2K) HTC Heat transfer coefficient (W/m2K) ICE Internal combustion engine k Ratio of specific heats (1.40) LHV Lower heating value MEP Mean effective pressure (kPa) P Pressure (kPa) R Ratio of Connecting Rod Length to Stroke Residual gas fraction (-) fr s Stroke (m) Specific entropy (J/g*K) s
SOC T TDC U Up V Vc Vd
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Start of combustion Temperature ( K) Top dead center Gas speed (cm/s) Mean piston speed (m/s) Volume (m3) Clearance volume (m3) Displacement volume (m3)
Subscripts and superscripts 1 State Point 1 of otto cycle (intake) 2 State point 2 of otto cycle (compression) 3 State point 3 of otto cycle (maximum value) 4 State point 4 of otto cycle (exhaust stroke) Greek symbols q Degrees of crank angle ( C.A) Dq Combustion duration ( C.A) f Equivalence ratio (-)
references
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