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Controlled auto-ignition characteristics of methane–air mixture in a rapid intake compression and expansion machine
Energy 35 (2010) 4184e4191
Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
Controlled auto-ignition characteristics of methaneeair mixture in a rapid intake compression and expansion machine Gyubaek Cho a, Dongsoo Jeong a, Gunfeel Moon b, Choongsik Bae c, * a
Engine Research Team, Eco-Machinery Research Division, Korea Institute of Machinery & Materials, 104 Sinseongno, Yuseong-gu, Daejeon 305-701, Republic of Korea Department of Clean Environmental system, University of Science & Technology, 52 Eoeun-dong, Yuseong-gu, Daejeon, Republic of Korea c Engine Laboratory, Department of Mechanical Engineering, Korea Advanced Institute of Science and Technology, 373-1 GuSeong-Dong, Yuseong-Gu, Daejeon 305-701, Republic of Korea b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 23 September 2009 Received in revised form 5 July 2010 Accepted 7 July 2010 Available online 13 August 2010
The characteristics of controlled auto-ignition (CAI) were investigated with a methaneeair mixture and simulated residual gas, that represents internal exhaust gas recirculation (IEGR). Supply systems were additionally installed on the conventional rapid compression machine (RCM), and this modified machineda rapid intake compression and expansion machine (RICEM)dwas able to simulate an intake stroke for the evaluation of controlled auto-ignition with fueleair mixture. The fueleair mixture and the simulated residual gas were introduced separately into the combustion chamber through the spool valves. Various IEGR rates and temperatures of the IEGR gas were tested. The initial reaction and the development in controlled auto-ignition combustion were compared with spark-ignited combustion by visualization with a high-speed digital camera. Under the controlled auto-ignition operation, multi-point ignition and faster combustion were observed. With increasing the temperature of IEGR gas, the autoignition timing was advanced and burning duration was shortened. The higher rate of IEGR had the same effects on the combustion of the controlled auto-ignition. However, this trend was reversed with more than 47 per cent of IEGR. Ó 2010 Elsevier Ltd. All rights reserved.
Keywords: Controlled auto-ignition (CAI) Methane Rapid intake compression and expansion machine (RICEM) Simulated residual gas Internal exhaust gas recirculation (IEGR)
1. Introduction Since the results of controlled auto-ignition (CAI) combustion, one of the homogeneous charge compression ignition (HCCI) technologies, in a two-stroke engine were first published in 1979 [1], many cases of application of CAI in four-stroke engines have been reported from the perspectives of numerical simulation and experimental research [2e5]. Nowadays, CAI is considered one of the most promising alternative combustion methods, which could replace conventional spark ignition (SI) for future engine technology. The CAI can be considered as a specific case of the HCCI application with high octane-number fuel, usually used in conventional SI engine, and achieved with residual gas trapping or hot exhaust gas recirculation (EGR) to enhance auto-ignition of the fuel. In order to achieve desirable CAI operation in all the range of engine operating conditions, the effects of intake air temperature [6,7], compression ratio [6,8], EGR rate [9e11], and fuel additives or
* Corresponding author. Tel.: þ82 42 350 3044; fax: þ82 42 350 5044. E-mail address: [email protected] (C. Bae). 0360-5442/$ e see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2010.07.002
compositions [12e17] have been investigated. In addition, many researchers have reported the characteristics of CAI combustion, including auto-ignition timing and temperature, combustion duration, and heat release rate [6e9,13,14]. Unfortunately, CAI combustion was restricted by excessive rate of in-cylinder pressure-rise, which can give rise to a ringing noise (knock) and may affect engine durability [18]. In single-cylinder or multi-cylinder engines, it is very difficult to evaluate the effect of each parameter on auto-ignition of the fueleair mixture or control of combustion, because parameters are coupled with each other under engine operating conditions. For this reason, some researchers have used a rapid compression machine (RCM) to elucidate auto-ignition characteristics and visualize combustion phenomena [4,19e22]. The RCM was designed to simulate the thermodynamic cycle of an engine, particularly compression and expansion strokes. The test results showed that the initial temperature and composition of fuel are the most important parameters that influence auto-ignition timing and extent of complete combustion. However, there is a restriction on the realization of intake and exhaust flow from the engine in the case of the RCM. Therefore, it is impossible not only to verify the effect of these flows on
G. Cho et al. / Energy 35 (2010) 4184e4191
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Fig. 1. Experimental apparatus.
combustion but also to evaluate the effect of the residual gas and EGR. In general, CAI combustion utilizes the heat from the residual gas (which is also called IEGR; internal exhaust gas recirculation) or from the external EGR to auto-ignite the fueleair mixture at normal compression ratios. CAI is also possible from the initiation of combustion induced by heating the intake mixture or by increasing the compression ratio without the help of the residual gas [23e25]. The authors evaluated the effects of IEGR conditions, such as supplying timing, homogeneity and equivalence ratio, in previous research [22]. In this research, previous results were expanded into the effects of other IEGR conditions on CAI. The combustion characteristics were analysed with respect to IEGR gas temperatures and IEGR rates using a rapid intake compression expansion machine (RICEM). In addition, the location of the initial reaction zone and the reaction development process in CAI combustion were investigated through direct imaging of the reaction zone.
2. Experimental apparatus and method The experimental setup consists of three parts: the RICEM system, the supply system of fueleair mixture and IEGR gas, and the data acquisition system [22]. The schematic diagram of the experimental setup is shown in Fig. 1. The RICEM system consists of the upper and lower cylinders, which are interconnected as shown in Fig. 2. The upper cylinder was driven by the movement of the lower cylinder to simulate a single cycle of the engine. The lower cylinder was operated by compressed air and had only an expansion function. The specifications of the RICEM are shown in Table 1. The operating procedure is as follows (refer to Fig. 3); 1) The pistons in the upper and lower cylinders were initially located at top dead centre (TDC). The lower piston was anchored by the pneumatic brake at the edge of the crank shaft and compressed air at 5 bar was charged in the lower cylinder during the compressed air charge process. 2) As soon as the pneumatic brake was released by the trigger signal, the lower piston started moving down to bottom dead centre (BDC), 360 crank angle degree (CAD), and subsequently caused the upper piston to move through the gear connection with a 2:1 gear ratio. 3) When the lower piston reached BDC, the upper piston passed BDC (intake stroke, 0e180 CAD) and reached TDC (compression stroke, 180e360 CAD). The IEGR gas and the fueleair mixture were inducted together into the combustion chamber during the intake stroke, and were subsequently mixed and compressed during the compression stroke.
Table 1 Specifications of rapid intake compression and expansion machine.
Fig. 2. Schematic diagram of the rapid intake, compression expansion machine (RICEM).
Item
Value
Bore Stroke (mm) Displacement (cc) Compression ratio Intake valve opening timing (ATDC) Exhaust valve opening timing (ATDC)
100 110 864 11.7 40e70 75e250
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Fig. 3. Operating mechanisms of a rapid intake, compression and expansion machine (RICEM).
4) Finally, the mixture started burning when the auto-ignition condition was reached. A spark plug was installed in the IEGR gas chamber to burn the fueleair mixture generating the IEGR gas. The IEGR gas and fueleair mixture chambers were mounted on the cylinder head and connected to the combustion chamber through spool-type valves, as shown in Fig. 4a. The spool-type valve was operated by pneumatic cylinder and was designed for opening and closing at one stroke. Methane (99.9% pure) and air came from each high pressure gas cylinder (Max. 120 bar, 47 L) and were premixed in a mixing tank at 10 bar. The fueleair mixture was supplied to the fueleair mixture chamber and the IEGR gas chamber at 2 bar. The equivalence ratio was controlled by adjusting the partial pressure of methane and air with a fine pressure gauge (Dwyer Instruments INC., DPGA-09).
Three fans were installed inside the mixing tank to guarantee homogeneous mixing. The combustion, the fueleair mixture, and the IEGR gas chambers were heated at 353 K (general temperature of water coolant) by an electric heater wrapped around the outer surface of the chambers for simulating real engine condition. The wall temperature of each chamber was measured by inserting thermocouple into the hole of the wall and controlled up to 1 accuracy by adjusting the voltage of electric heater power. For calculating the amounts of air-fuel mixture and IEGR incylinder and the IEGR rate, the pressures in the cylinder and in the two chambers were measured with piezoelectric pressure transducers (Kistler, 6117BFD and 6067C1). The sampling frequencies of the pressures and encoder (Omron, E6D-CWZ1E, 3600 puls/rev.) signals were 10 kHz. The direct images were acquired at a rate of 2000 frames per second by high-speed digital camera (Phantom co. Ltd., V7.3) through the quartz piston head and a 45-degree mirror in the RICEM, as shown in Fig. 4b. Table 2 shows the accuracy of the measurements and the uncertainty of the estimated results of the various parameters. The experimental conditions of this study are summarized in Table 3. The combustion chamber temperature was fixed and other parameters, such as IEGR rate and the temperature of IEGR gas, were varied independently to investigate their effects on CAI combustion.
Table 2 Accuracy of measurements and uncertainty of estimated results. Measurements
Accuracy
Pressure (Kistler 6067C1) Pressure (Kistler 6117BFD) Pressure (Dwyer DPGA-09) Temperature (K-type) Crank shaft encoder
1% 3% 0.005 bar 2.2 C 0.05 /pulse
Estimation results
Uncertainty (%)
IEGR rate Equivalence ratio
3.2 0.05
Table 3 Test conditions.
Fig. 4. Structure of cylinder head and piston (a) Cylinder head (b) Elongated piston structure and visualization area.
Parameter
Condition
Combustion chamber temperature (K) IEGR gas supplying timing (ATDC) Equivalence ratio (F) IEGR gas temperature (K) IEGR rate (per cent)
353 100e140 1.0 1000e1300 25e65
G. Cho et al. / Energy 35 (2010) 4184e4191
Hc;iþ1 ¼ Hc;i þ Hm;i þ Hb;i
3. Results and discussion The mass of the fueleair mixture was calculated from the state equation of an ideal gas with the pressure difference between before valve opening and after mixture supply [16,26]. The mass flow rate of the IEGR gas can be calculated as subcritical or critical gas flow, based on whether the pressure ratio of IEGR gas chamber to the combustion chamber was more than 0.532 or less. In subcritical gas flow, the mass flow rate of the IEGR gas was obtained from the flow equation through a restriction [26]. In case of critical gas flow, the flow is choked and equation (1) was used.
ðgþ1Þ=ð2ðg1ÞÞ Cd AT Pb 1=2 2 _ ¼ p ffiffiffiffiffiffiffiffi g m gþ1 RT0
(1)
_ is the mass flow rate, Cd is the flow coefficient, AT is the where m valve area, Pb is the IEGR chamber pressure, R is the gas constant, T0 is the initial gas temperature, and g is the specific heat ratio. The IEGR rate was defined as the ratio of the mass of the IEGR gas, integrating mass flow rate (eq. (1)) during the exhaust valve open, to the total mass in chamber as follows;
IEGR rateð%Þ ¼
mb 100 mm þ mb
4187
(2)
where mb and mm are the mass of the IEGR gas and the fueleair mixture, respectively. In order to calculate the in-cylinder gas temperature, different assumptions were used for each period. For the expansion and the compression periods, the adiabatic process was assumed. The expansion period is between TDC (start of piston movement) and ‘valve open’, whereas the compression period is defined as the time between ‘valve close’ and the start of combustion. Therefore, the temperature of the in-cylinder gas was calculated using isentropic relations [26]. For the gas introduction period, the in-cylinder gas temperature was calculated using the enthalpy balance between the fueleair mixtures, the IEGR gas, and the in-cylinder gas in the combustion chamber [27].
Tc;iþ1 ¼
(3)
mc;i cp;cðiÞ Tc;i þ mm;i cp;mðiÞ Tm;i þ mb;i cp;bðiÞ Tb;i mc;i cp;cðiÞ þ mm;i cp;mðiÞ þ mb;i cp;bðiÞ
where cp;cðiÞ ¼
(4)
mc;i1 cp;cði1Þ þ mm;i1 cp;mði1Þ þ mb;i1 cp;bði1Þ mc;i1 þ mm;i1 þ mb;i1
Nomenclature: H: enthalpy of gas, T: gas temperature, m: mass of gas, cp : specific heat capacity at constant pressure. Subscripts: c: in-cylinder gas, m: fueleair mixture, b: IEGR gas, i 1; i; i þ 1: index. The start of ignition (SOI) and end of combustion (EOC) were defined as the timing when the heat release rate (HRR) became greater and less than 1 J per CAD, respectively [26].
g dV dQn dQloss 1 dP ¼ þ þ P V g 1 dq g 1 dq dq dq
(5)
where Qn is the net heat release and q is the crank angle. Qloss is the heat loss to the wall and is ignored in this research. Cylinder head is not cooled by coolant but heated by electric heater rather for keeping operation condition. Finally, the combustion efficiency was defined as the ratio of the total amount of heat release to the lower heating value (LHV) of the fuel [16]. The combustion efficiency depends on the balance of loss and work in a cycle and the optimum auto-ignition timing was identified at the highest combustion efficiency among the test conditions
hc ¼
P
HRR 100ð%Þ LHV mf
(6)
where mf is the mass of fuel. 3.1. Visualization of reaction fronts The initial reaction regions of CAI combustion taken as a single shot image at different cycles in the RICEM are presented in Fig. 5.
Fig. 5. Reaction regions of controlled auto-ignition at different cycles. (F ¼ 1.0, rc ¼ 11.7).
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Fig. 6. Reaction front development of controlled auto-ignition combustion in one cycle. (F ¼ 1.0, rc ¼ 11.7).
The locations of reaction initiation are random and the shapes of the reaction region are irregular. These images of the reaction region in the RICEM verify that spontaneous ignitions were initiated and reaction region developed at multiple points at the same time, which is a typical feature of CAI combustion. The strength of the reaction region was weaker than the flame of SI combustion due to the lower temperature of the reaction [19]. Fig. 6 shows an example of the reaction front development process and the pressure curve when auto-ignition occurred due to the compression of the fueleair mixture with the IEGR gas. More
than three different auto-ignited points were observed at around 3 CAD before top dead centre (BTDC), and the reaction regions rapidly developed from the centre of the ignition point to the unburned gas zone. The combustion duration was about 7 CAD. The typical flame propagation in SI combustion is shown in Fig. 7. In this figure, the flame kernel is formed around the spark plug (the centre of the flame image) around BTDC 4 CAD and propagates slowly across the combustion chamber. Comparing the characteristics of CAI and SI combustion, the location of peak pressure is advanced and the value of the peak
Fig. 7. Flame propagation of spark ignition combustion in one cycle. (F ¼ 1.0, tig ¼ BTDC 4 deg. rc ¼ 11.7).
G. Cho et al. / Energy 35 (2010) 4184e4191
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Fig. 8. Cylinder combustion pressure and heat release rate for different IEGR gas temperatures. (F ¼ 1.0, rc ¼ 11.7, EGR rate ¼ 30%, tb ¼ 140 CAD).
Fig. 10. Cylinder combustion pressure and heat release rate for different IEGR rates. (F ¼ 1.0, rc ¼ 11.7, EGR rate: 27e35%, Tb ¼ 1200 K, tb ¼ 120 CAD).
pressure is lowered in CAI combustion. The reaction process of CAI is mainly dominated by the low-temperature chemistry because CAI combustion is initiated by auto-ignition at relatively lowtemperature. Even though the speed of reaction front development in CAI combustion is much faster than that in SI combustion, the
burned gas temperature is much lower due to its inactive components. This is why the peak pressure in CAI combustion is lower than that in SI combustion. 3.2. The effect of IEGR gas temperature Fig. 8 illustrates the pressure and the HRR for various IEGR temperatures. The IEGR rate was fixed at 30%, and the IEGR supply timing was also fixed at 140 CAD at the stoichiometric condition with the compression ratio of 11.7. The auto-ignition timing was advanced as the temperature of the IEGR gas was increased. The IEGR gas was burned gas at high temperature which helped auto-ignition of fueleair mixture. Therefore, the higher temperature of IEGR gas gives higher thermal energy and causes earlier auto-ignition timing. This effect is considered as ‘the thermal effect’ and is similar to the inlet charge preheating method. The auto-ignition timing was advanced by about 3.3 CAD as the temperature of the IEGR gas was increased by about 100 K. Fig. 9 shows the timings for SOI, EOC, maximum HRR, and the relationship between the maximum HRR, total heat release amount, combustion efficiency, and IEGR gas temperature. The auto-ignition timing was advanced with higher temperature of the IEGR gas, whereas the timing of the maximum HRR was retarded
Fig. 9. Combustion characteristics for different IEGR gas temperatures. (a) Start and end of combustion timing. (b) Maximum rate of heat release, heat release and combustion efficiency. (F ¼ 1.0, rc ¼ 11.7, EGR rate ¼ 30%, tb ¼ 140 CAD).
Fig. 11. Cylinder combustion pressure and heat release rate for different IEGR rates. (F ¼ 1.0, rc ¼ 11.7, EGR rate: 47e65%: Tb ¼ 1150 K, tb ¼ 100 CAD).
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Fig. 14. Maximum rate of heat release, heat release and combustion efficiency for different IEGR rate. Fig. 12. Start and end of combustion timing for different IEGR rates. (F ¼ 1.0, rc ¼ 11.7, EGR rate: 27e35%: Tb ¼ 1200 K, tb ¼ 120 CAD, EGR rate: 47e65%: Tb ¼ 1150 K, tb ¼ 100 CAD).
and occurred closer to the EOC. This means that it takes a certain amount of time for the flame to propagate towards the mixture after auto-ignition. Optimum auto-ignition takes place at 1077 K with the highest combustion efficiency, while the combustion efficiency varied slightly with the temperature at around 40%.
3.3. The effect of the IEGR rate The variations of the pressure and the heat release rate with respect to IEGR rate are shown in Fig.10. The IEGR rate was controlled in the range of 25e65% by adjusting the exhaust valve timing and valve geometry. The IEGR gas supply timing of 120 CAD and the IEGR gas temperature of 1200 K were utilized for IEGR rate of 27e35% at stoichiometric conditions with a compression ratio of 11.7. Fig. 10 indicates that the auto-ignition timing was advanced as the IEGR rate increased. The advancement of the auto-ignition timing was due to the rise of the fueleair mixture temperature caused by the increased amount of the IEGR gas at high temperature. Eventually, the increased IEGR rate shows the same effect as the increased temperature of the IEGR gas. That is, a 7.5% increase in
IEGR rate advances the auto-ignition angle by 7 CAD, which corresponds to the effect of a 200 K temperature increase in the IEGR gas. The IEGR gas supply timing of 100 CAD, and the IEGR gas temperature of 1150 K were used for the IEGR rate of 47e65% at the same compression ratio and equivalence ratio as the previous conditions in Fig. 11. In contrast to previous results, auto-ignition timing was delayed as the IEGR rate increased. Fig. 12 indicates the SOI and the EOC calculated from the conditions shown in Figs. 10 and 11. Although the auto-ignition timing was advanced and the burn duration was shortened as the IEGR rate was increased up to 47%, the auto-ignition timing was retarded and the combustion duration remained relatively constant for an EGR rate of more than 47%. This is attributed to the increases of the specific heat ratio, heat loss, and combustion instability due to the reduction in the oxygen concentration in the combustion chamber (which is also called the dilution effect of IEGR). It means that the dilution effect of IEGR gas is higher than thermal effect in high IEGR rate conditions. Fig. 13 shows the relation of the auto-ignition temperature versus the IEGR rate. In this graph, the auto-ignition temperature was calculated at the SOI using isentropic relations and eq. (4). As the IEGR rate was increased, the temperature of the auto-ignition was also increased due to the dilution of fueleair mixture by IEGR gas and the increment of specific heat capacity. Fig. 14 shows the relationship between the maximum HRR, total heat release amount, combustion efficiency and IEGR rate. The higher IEGR rate has lower combustion efficiency due to the increase of crevice loss and blow-by with higher pressure-rise rate by earlier auto-ignition before TDC. But lower IEGR rate than 27.1% showed the misfire of fueleair mixture frequently. Therefore, 27.1% of IEGR rate was found as the best condition in this research. 4. Conclusions
Fig. 13. Auto-ignition temperatures for different IEGR rates.
The quantitative evaluation of controlled auto-ignition (CAI) combustion was performed with the rapid intake compression and expansion machine (RICEM) by varying the temperature of simulated residual gas, that represents internal exhaust gas recirculation (IEGR), and IEGR rate. The following conclusions were drawn from the experimental results. The characteristics of CAI and spark ignition (SI) combustion in the RICEM were examined through comparison of direct reaction images and pressure traces in the combustion chamber. The speed
G. Cho et al. / Energy 35 (2010) 4184e4191
of reaction zone development in CAI combustion is much faster than that in SI combustion. The higher the temperature of the IEGR, the earlier auto-ignition timing occurred and the shorter the combustion duration due to thermal effect. Quantitatively, auto-ignition timing was advanced by 3.3 Crank angle degree (CAD) as the IEGR temperature was increased by 100 K from 1100 K to 1300 K. The auto-ignition timing was advanced and combustion duration was shortened as the IEGR rate increased up to 47% due to thermal effect by IEGR gas. However, the auto-ignition timing was rather delayed at the IEGR rate more than 47%. The excessively supplied IEGR gas (47%e65%) showed detrimental influence on the auto-ignition of fuel because of the less flammability by the dilution effect and the heat absorption caused by the high specific heat ratio of IEGR gas.
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Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
Controlled auto-ignition characteristics of methaneeair mixture in a rapid intake compression and expansion machine Gyubaek Cho a, Dongsoo Jeong a, Gunfeel Moon b, Choongsik Bae c, * a
Engine Research Team, Eco-Machinery Research Division, Korea Institute of Machinery & Materials, 104 Sinseongno, Yuseong-gu, Daejeon 305-701, Republic of Korea Department of Clean Environmental system, University of Science & Technology, 52 Eoeun-dong, Yuseong-gu, Daejeon, Republic of Korea c Engine Laboratory, Department of Mechanical Engineering, Korea Advanced Institute of Science and Technology, 373-1 GuSeong-Dong, Yuseong-Gu, Daejeon 305-701, Republic of Korea b
a r t i c l e i n f o
a b s t r a c t
Article history: Received 23 September 2009 Received in revised form 5 July 2010 Accepted 7 July 2010 Available online 13 August 2010
The characteristics of controlled auto-ignition (CAI) were investigated with a methaneeair mixture and simulated residual gas, that represents internal exhaust gas recirculation (IEGR). Supply systems were additionally installed on the conventional rapid compression machine (RCM), and this modified machineda rapid intake compression and expansion machine (RICEM)dwas able to simulate an intake stroke for the evaluation of controlled auto-ignition with fueleair mixture. The fueleair mixture and the simulated residual gas were introduced separately into the combustion chamber through the spool valves. Various IEGR rates and temperatures of the IEGR gas were tested. The initial reaction and the development in controlled auto-ignition combustion were compared with spark-ignited combustion by visualization with a high-speed digital camera. Under the controlled auto-ignition operation, multi-point ignition and faster combustion were observed. With increasing the temperature of IEGR gas, the autoignition timing was advanced and burning duration was shortened. The higher rate of IEGR had the same effects on the combustion of the controlled auto-ignition. However, this trend was reversed with more than 47 per cent of IEGR. Ó 2010 Elsevier Ltd. All rights reserved.
Keywords: Controlled auto-ignition (CAI) Methane Rapid intake compression and expansion machine (RICEM) Simulated residual gas Internal exhaust gas recirculation (IEGR)
1. Introduction Since the results of controlled auto-ignition (CAI) combustion, one of the homogeneous charge compression ignition (HCCI) technologies, in a two-stroke engine were first published in 1979 [1], many cases of application of CAI in four-stroke engines have been reported from the perspectives of numerical simulation and experimental research [2e5]. Nowadays, CAI is considered one of the most promising alternative combustion methods, which could replace conventional spark ignition (SI) for future engine technology. The CAI can be considered as a specific case of the HCCI application with high octane-number fuel, usually used in conventional SI engine, and achieved with residual gas trapping or hot exhaust gas recirculation (EGR) to enhance auto-ignition of the fuel. In order to achieve desirable CAI operation in all the range of engine operating conditions, the effects of intake air temperature [6,7], compression ratio [6,8], EGR rate [9e11], and fuel additives or
* Corresponding author. Tel.: þ82 42 350 3044; fax: þ82 42 350 5044. E-mail address: [email protected] (C. Bae). 0360-5442/$ e see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2010.07.002
compositions [12e17] have been investigated. In addition, many researchers have reported the characteristics of CAI combustion, including auto-ignition timing and temperature, combustion duration, and heat release rate [6e9,13,14]. Unfortunately, CAI combustion was restricted by excessive rate of in-cylinder pressure-rise, which can give rise to a ringing noise (knock) and may affect engine durability [18]. In single-cylinder or multi-cylinder engines, it is very difficult to evaluate the effect of each parameter on auto-ignition of the fueleair mixture or control of combustion, because parameters are coupled with each other under engine operating conditions. For this reason, some researchers have used a rapid compression machine (RCM) to elucidate auto-ignition characteristics and visualize combustion phenomena [4,19e22]. The RCM was designed to simulate the thermodynamic cycle of an engine, particularly compression and expansion strokes. The test results showed that the initial temperature and composition of fuel are the most important parameters that influence auto-ignition timing and extent of complete combustion. However, there is a restriction on the realization of intake and exhaust flow from the engine in the case of the RCM. Therefore, it is impossible not only to verify the effect of these flows on
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Fig. 1. Experimental apparatus.
combustion but also to evaluate the effect of the residual gas and EGR. In general, CAI combustion utilizes the heat from the residual gas (which is also called IEGR; internal exhaust gas recirculation) or from the external EGR to auto-ignite the fueleair mixture at normal compression ratios. CAI is also possible from the initiation of combustion induced by heating the intake mixture or by increasing the compression ratio without the help of the residual gas [23e25]. The authors evaluated the effects of IEGR conditions, such as supplying timing, homogeneity and equivalence ratio, in previous research [22]. In this research, previous results were expanded into the effects of other IEGR conditions on CAI. The combustion characteristics were analysed with respect to IEGR gas temperatures and IEGR rates using a rapid intake compression expansion machine (RICEM). In addition, the location of the initial reaction zone and the reaction development process in CAI combustion were investigated through direct imaging of the reaction zone.
2. Experimental apparatus and method The experimental setup consists of three parts: the RICEM system, the supply system of fueleair mixture and IEGR gas, and the data acquisition system [22]. The schematic diagram of the experimental setup is shown in Fig. 1. The RICEM system consists of the upper and lower cylinders, which are interconnected as shown in Fig. 2. The upper cylinder was driven by the movement of the lower cylinder to simulate a single cycle of the engine. The lower cylinder was operated by compressed air and had only an expansion function. The specifications of the RICEM are shown in Table 1. The operating procedure is as follows (refer to Fig. 3); 1) The pistons in the upper and lower cylinders were initially located at top dead centre (TDC). The lower piston was anchored by the pneumatic brake at the edge of the crank shaft and compressed air at 5 bar was charged in the lower cylinder during the compressed air charge process. 2) As soon as the pneumatic brake was released by the trigger signal, the lower piston started moving down to bottom dead centre (BDC), 360 crank angle degree (CAD), and subsequently caused the upper piston to move through the gear connection with a 2:1 gear ratio. 3) When the lower piston reached BDC, the upper piston passed BDC (intake stroke, 0e180 CAD) and reached TDC (compression stroke, 180e360 CAD). The IEGR gas and the fueleair mixture were inducted together into the combustion chamber during the intake stroke, and were subsequently mixed and compressed during the compression stroke.
Table 1 Specifications of rapid intake compression and expansion machine.
Fig. 2. Schematic diagram of the rapid intake, compression expansion machine (RICEM).
Item
Value
Bore Stroke (mm) Displacement (cc) Compression ratio Intake valve opening timing (ATDC) Exhaust valve opening timing (ATDC)
100 110 864 11.7 40e70 75e250
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Fig. 3. Operating mechanisms of a rapid intake, compression and expansion machine (RICEM).
4) Finally, the mixture started burning when the auto-ignition condition was reached. A spark plug was installed in the IEGR gas chamber to burn the fueleair mixture generating the IEGR gas. The IEGR gas and fueleair mixture chambers were mounted on the cylinder head and connected to the combustion chamber through spool-type valves, as shown in Fig. 4a. The spool-type valve was operated by pneumatic cylinder and was designed for opening and closing at one stroke. Methane (99.9% pure) and air came from each high pressure gas cylinder (Max. 120 bar, 47 L) and were premixed in a mixing tank at 10 bar. The fueleair mixture was supplied to the fueleair mixture chamber and the IEGR gas chamber at 2 bar. The equivalence ratio was controlled by adjusting the partial pressure of methane and air with a fine pressure gauge (Dwyer Instruments INC., DPGA-09).
Three fans were installed inside the mixing tank to guarantee homogeneous mixing. The combustion, the fueleair mixture, and the IEGR gas chambers were heated at 353 K (general temperature of water coolant) by an electric heater wrapped around the outer surface of the chambers for simulating real engine condition. The wall temperature of each chamber was measured by inserting thermocouple into the hole of the wall and controlled up to 1 accuracy by adjusting the voltage of electric heater power. For calculating the amounts of air-fuel mixture and IEGR incylinder and the IEGR rate, the pressures in the cylinder and in the two chambers were measured with piezoelectric pressure transducers (Kistler, 6117BFD and 6067C1). The sampling frequencies of the pressures and encoder (Omron, E6D-CWZ1E, 3600 puls/rev.) signals were 10 kHz. The direct images were acquired at a rate of 2000 frames per second by high-speed digital camera (Phantom co. Ltd., V7.3) through the quartz piston head and a 45-degree mirror in the RICEM, as shown in Fig. 4b. Table 2 shows the accuracy of the measurements and the uncertainty of the estimated results of the various parameters. The experimental conditions of this study are summarized in Table 3. The combustion chamber temperature was fixed and other parameters, such as IEGR rate and the temperature of IEGR gas, were varied independently to investigate their effects on CAI combustion.
Table 2 Accuracy of measurements and uncertainty of estimated results. Measurements
Accuracy
Pressure (Kistler 6067C1) Pressure (Kistler 6117BFD) Pressure (Dwyer DPGA-09) Temperature (K-type) Crank shaft encoder
1% 3% 0.005 bar 2.2 C 0.05 /pulse
Estimation results
Uncertainty (%)
IEGR rate Equivalence ratio
3.2 0.05
Table 3 Test conditions.
Fig. 4. Structure of cylinder head and piston (a) Cylinder head (b) Elongated piston structure and visualization area.
Parameter
Condition
Combustion chamber temperature (K) IEGR gas supplying timing (ATDC) Equivalence ratio (F) IEGR gas temperature (K) IEGR rate (per cent)
353 100e140 1.0 1000e1300 25e65
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Hc;iþ1 ¼ Hc;i þ Hm;i þ Hb;i
3. Results and discussion The mass of the fueleair mixture was calculated from the state equation of an ideal gas with the pressure difference between before valve opening and after mixture supply [16,26]. The mass flow rate of the IEGR gas can be calculated as subcritical or critical gas flow, based on whether the pressure ratio of IEGR gas chamber to the combustion chamber was more than 0.532 or less. In subcritical gas flow, the mass flow rate of the IEGR gas was obtained from the flow equation through a restriction [26]. In case of critical gas flow, the flow is choked and equation (1) was used.
ðgþ1Þ=ð2ðg1ÞÞ Cd AT Pb 1=2 2 _ ¼ p ffiffiffiffiffiffiffiffi g m gþ1 RT0
(1)
_ is the mass flow rate, Cd is the flow coefficient, AT is the where m valve area, Pb is the IEGR chamber pressure, R is the gas constant, T0 is the initial gas temperature, and g is the specific heat ratio. The IEGR rate was defined as the ratio of the mass of the IEGR gas, integrating mass flow rate (eq. (1)) during the exhaust valve open, to the total mass in chamber as follows;
IEGR rateð%Þ ¼
mb 100 mm þ mb
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(2)
where mb and mm are the mass of the IEGR gas and the fueleair mixture, respectively. In order to calculate the in-cylinder gas temperature, different assumptions were used for each period. For the expansion and the compression periods, the adiabatic process was assumed. The expansion period is between TDC (start of piston movement) and ‘valve open’, whereas the compression period is defined as the time between ‘valve close’ and the start of combustion. Therefore, the temperature of the in-cylinder gas was calculated using isentropic relations [26]. For the gas introduction period, the in-cylinder gas temperature was calculated using the enthalpy balance between the fueleair mixtures, the IEGR gas, and the in-cylinder gas in the combustion chamber [27].
Tc;iþ1 ¼
(3)
mc;i cp;cðiÞ Tc;i þ mm;i cp;mðiÞ Tm;i þ mb;i cp;bðiÞ Tb;i mc;i cp;cðiÞ þ mm;i cp;mðiÞ þ mb;i cp;bðiÞ
where cp;cðiÞ ¼
(4)
mc;i1 cp;cði1Þ þ mm;i1 cp;mði1Þ þ mb;i1 cp;bði1Þ mc;i1 þ mm;i1 þ mb;i1
Nomenclature: H: enthalpy of gas, T: gas temperature, m: mass of gas, cp : specific heat capacity at constant pressure. Subscripts: c: in-cylinder gas, m: fueleair mixture, b: IEGR gas, i 1; i; i þ 1: index. The start of ignition (SOI) and end of combustion (EOC) were defined as the timing when the heat release rate (HRR) became greater and less than 1 J per CAD, respectively [26].
g dV dQn dQloss 1 dP ¼ þ þ P V g 1 dq g 1 dq dq dq
(5)
where Qn is the net heat release and q is the crank angle. Qloss is the heat loss to the wall and is ignored in this research. Cylinder head is not cooled by coolant but heated by electric heater rather for keeping operation condition. Finally, the combustion efficiency was defined as the ratio of the total amount of heat release to the lower heating value (LHV) of the fuel [16]. The combustion efficiency depends on the balance of loss and work in a cycle and the optimum auto-ignition timing was identified at the highest combustion efficiency among the test conditions
hc ¼
P
HRR 100ð%Þ LHV mf
(6)
where mf is the mass of fuel. 3.1. Visualization of reaction fronts The initial reaction regions of CAI combustion taken as a single shot image at different cycles in the RICEM are presented in Fig. 5.
Fig. 5. Reaction regions of controlled auto-ignition at different cycles. (F ¼ 1.0, rc ¼ 11.7).
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Fig. 6. Reaction front development of controlled auto-ignition combustion in one cycle. (F ¼ 1.0, rc ¼ 11.7).
The locations of reaction initiation are random and the shapes of the reaction region are irregular. These images of the reaction region in the RICEM verify that spontaneous ignitions were initiated and reaction region developed at multiple points at the same time, which is a typical feature of CAI combustion. The strength of the reaction region was weaker than the flame of SI combustion due to the lower temperature of the reaction [19]. Fig. 6 shows an example of the reaction front development process and the pressure curve when auto-ignition occurred due to the compression of the fueleair mixture with the IEGR gas. More
than three different auto-ignited points were observed at around 3 CAD before top dead centre (BTDC), and the reaction regions rapidly developed from the centre of the ignition point to the unburned gas zone. The combustion duration was about 7 CAD. The typical flame propagation in SI combustion is shown in Fig. 7. In this figure, the flame kernel is formed around the spark plug (the centre of the flame image) around BTDC 4 CAD and propagates slowly across the combustion chamber. Comparing the characteristics of CAI and SI combustion, the location of peak pressure is advanced and the value of the peak
Fig. 7. Flame propagation of spark ignition combustion in one cycle. (F ¼ 1.0, tig ¼ BTDC 4 deg. rc ¼ 11.7).
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Fig. 8. Cylinder combustion pressure and heat release rate for different IEGR gas temperatures. (F ¼ 1.0, rc ¼ 11.7, EGR rate ¼ 30%, tb ¼ 140 CAD).
Fig. 10. Cylinder combustion pressure and heat release rate for different IEGR rates. (F ¼ 1.0, rc ¼ 11.7, EGR rate: 27e35%, Tb ¼ 1200 K, tb ¼ 120 CAD).
pressure is lowered in CAI combustion. The reaction process of CAI is mainly dominated by the low-temperature chemistry because CAI combustion is initiated by auto-ignition at relatively lowtemperature. Even though the speed of reaction front development in CAI combustion is much faster than that in SI combustion, the
burned gas temperature is much lower due to its inactive components. This is why the peak pressure in CAI combustion is lower than that in SI combustion. 3.2. The effect of IEGR gas temperature Fig. 8 illustrates the pressure and the HRR for various IEGR temperatures. The IEGR rate was fixed at 30%, and the IEGR supply timing was also fixed at 140 CAD at the stoichiometric condition with the compression ratio of 11.7. The auto-ignition timing was advanced as the temperature of the IEGR gas was increased. The IEGR gas was burned gas at high temperature which helped auto-ignition of fueleair mixture. Therefore, the higher temperature of IEGR gas gives higher thermal energy and causes earlier auto-ignition timing. This effect is considered as ‘the thermal effect’ and is similar to the inlet charge preheating method. The auto-ignition timing was advanced by about 3.3 CAD as the temperature of the IEGR gas was increased by about 100 K. Fig. 9 shows the timings for SOI, EOC, maximum HRR, and the relationship between the maximum HRR, total heat release amount, combustion efficiency, and IEGR gas temperature. The auto-ignition timing was advanced with higher temperature of the IEGR gas, whereas the timing of the maximum HRR was retarded
Fig. 9. Combustion characteristics for different IEGR gas temperatures. (a) Start and end of combustion timing. (b) Maximum rate of heat release, heat release and combustion efficiency. (F ¼ 1.0, rc ¼ 11.7, EGR rate ¼ 30%, tb ¼ 140 CAD).
Fig. 11. Cylinder combustion pressure and heat release rate for different IEGR rates. (F ¼ 1.0, rc ¼ 11.7, EGR rate: 47e65%: Tb ¼ 1150 K, tb ¼ 100 CAD).
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Fig. 14. Maximum rate of heat release, heat release and combustion efficiency for different IEGR rate. Fig. 12. Start and end of combustion timing for different IEGR rates. (F ¼ 1.0, rc ¼ 11.7, EGR rate: 27e35%: Tb ¼ 1200 K, tb ¼ 120 CAD, EGR rate: 47e65%: Tb ¼ 1150 K, tb ¼ 100 CAD).
and occurred closer to the EOC. This means that it takes a certain amount of time for the flame to propagate towards the mixture after auto-ignition. Optimum auto-ignition takes place at 1077 K with the highest combustion efficiency, while the combustion efficiency varied slightly with the temperature at around 40%.
3.3. The effect of the IEGR rate The variations of the pressure and the heat release rate with respect to IEGR rate are shown in Fig.10. The IEGR rate was controlled in the range of 25e65% by adjusting the exhaust valve timing and valve geometry. The IEGR gas supply timing of 120 CAD and the IEGR gas temperature of 1200 K were utilized for IEGR rate of 27e35% at stoichiometric conditions with a compression ratio of 11.7. Fig. 10 indicates that the auto-ignition timing was advanced as the IEGR rate increased. The advancement of the auto-ignition timing was due to the rise of the fueleair mixture temperature caused by the increased amount of the IEGR gas at high temperature. Eventually, the increased IEGR rate shows the same effect as the increased temperature of the IEGR gas. That is, a 7.5% increase in
IEGR rate advances the auto-ignition angle by 7 CAD, which corresponds to the effect of a 200 K temperature increase in the IEGR gas. The IEGR gas supply timing of 100 CAD, and the IEGR gas temperature of 1150 K were used for the IEGR rate of 47e65% at the same compression ratio and equivalence ratio as the previous conditions in Fig. 11. In contrast to previous results, auto-ignition timing was delayed as the IEGR rate increased. Fig. 12 indicates the SOI and the EOC calculated from the conditions shown in Figs. 10 and 11. Although the auto-ignition timing was advanced and the burn duration was shortened as the IEGR rate was increased up to 47%, the auto-ignition timing was retarded and the combustion duration remained relatively constant for an EGR rate of more than 47%. This is attributed to the increases of the specific heat ratio, heat loss, and combustion instability due to the reduction in the oxygen concentration in the combustion chamber (which is also called the dilution effect of IEGR). It means that the dilution effect of IEGR gas is higher than thermal effect in high IEGR rate conditions. Fig. 13 shows the relation of the auto-ignition temperature versus the IEGR rate. In this graph, the auto-ignition temperature was calculated at the SOI using isentropic relations and eq. (4). As the IEGR rate was increased, the temperature of the auto-ignition was also increased due to the dilution of fueleair mixture by IEGR gas and the increment of specific heat capacity. Fig. 14 shows the relationship between the maximum HRR, total heat release amount, combustion efficiency and IEGR rate. The higher IEGR rate has lower combustion efficiency due to the increase of crevice loss and blow-by with higher pressure-rise rate by earlier auto-ignition before TDC. But lower IEGR rate than 27.1% showed the misfire of fueleair mixture frequently. Therefore, 27.1% of IEGR rate was found as the best condition in this research. 4. Conclusions
Fig. 13. Auto-ignition temperatures for different IEGR rates.
The quantitative evaluation of controlled auto-ignition (CAI) combustion was performed with the rapid intake compression and expansion machine (RICEM) by varying the temperature of simulated residual gas, that represents internal exhaust gas recirculation (IEGR), and IEGR rate. The following conclusions were drawn from the experimental results. The characteristics of CAI and spark ignition (SI) combustion in the RICEM were examined through comparison of direct reaction images and pressure traces in the combustion chamber. The speed
G. Cho et al. / Energy 35 (2010) 4184e4191
of reaction zone development in CAI combustion is much faster than that in SI combustion. The higher the temperature of the IEGR, the earlier auto-ignition timing occurred and the shorter the combustion duration due to thermal effect. Quantitatively, auto-ignition timing was advanced by 3.3 Crank angle degree (CAD) as the IEGR temperature was increased by 100 K from 1100 K to 1300 K. The auto-ignition timing was advanced and combustion duration was shortened as the IEGR rate increased up to 47% due to thermal effect by IEGR gas. However, the auto-ignition timing was rather delayed at the IEGR rate more than 47%. The excessively supplied IEGR gas (47%e65%) showed detrimental influence on the auto-ignition of fuel because of the less flammability by the dilution effect and the heat absorption caused by the high specific heat ratio of IEGR gas.
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