Gasoline−Di-methyl Ether Homogeneous Charge Compression

1942

Energy & Fuels 2007, 21, 1942-1949

Gasoline-Di-methyl Ether Homogeneous Charge Compression Ignition Engine Kitae Yeom and Choongsik Bae* Korea AdVanced Institute of Science and Technology (KAIST), Taejon 305-701, Republic of Korea ReceiVed February 12, 2007. ReVised Manuscript ReceiVed May 3, 2007

The combustion and knock characteristics in an engine were investigated under the homogeneous charge compression ignition (HCCI) operation condition fueled with and gasoline and di-methyl ether (DME) with regard to the valve timing, in order to identify an operating limit and correlation between knock and operating conditions. Commercial gasoline was used as fuel and injected at the intake port using port fuel injection equipment. Di-methyl ether (DME) was used as an ignition promoter and was injected directly into the cylinder during the intake stroke. Different intake valve timings and fuel amounts were tested to verify the knock characteristics of the HCCI engine. In order to characterize the high load operation limit, the peak combustion pressure, rate of combustion pressure rise, and indicated mean effective pressure were correlated with the knocking probability and intensity. The start of combustion was independent from the intake valve open (IVO) timing. The burn duration and the rate of the combustion pressure rise depended on the fuel quantity and the IVO timing. The knocking probability and intensity increased with the peak combustion pressure and the combustion pressure rise. According to the indicated mean effective pressure decrease when the total air excess ratio was lower than 2.4, the high load operating range of the test engine was limited to under 0.5 MW/m2. The knocking probability can be predicted by two derived equations. The hydrocarbon emission increased at the retarded intake valve open timing and the lower total air excess ratio. The carbon dioxide emission was influenced mainly by the total air excess ratio.

Introduction The main concepts of homogeneous charge compression ignition (HCCI) are the breathing of a premixed air-fuel mixture as in a spark ignition (SI) engine, and the ignition without a spark plug as in a compression ignition (CI) engine.2 In order to reduce nitrogen oxide (NOx) and particulate matter (PM) emission simultaneously, the HCCI combustion concept has been introduced.3 However, HCCI combustion leads to high hydrocarbon (HC) and carbon monoxide (CO) emissions. These high HC and CO emissions are the result of a low combustion temperature, which leads to low NOx emission. An ultralean mixture releases less heat than a rich mixture. HC and CO emissions are insufficiently oxidized due to a lower combustion temperature.2 However, CO emission originates from a lack of the oxidation reaction during the expansion stroke. The CO emission is an index of the incomplete combustion.4 Self-ignition of an HCCI engine originates from the surrounding air-fuel * Corresponding author. Professor, Engine Laboratory, Department of Mechanical Engineering, Korea Advanced Institute of Science and Technology, 373-1, Guseong-dong, Yuseong-gu, Taejon 305-701, Republic of Korea. Tel.: +82-42-869-3044. E-mail: [email protected]. (1) Chun, K.; Kim, K. Measurement and Analysis of Knock in a SI Engine Using the Cylinder Pressure and Block Vibration Signals. SAE Trans. 1994, 103 (3), 56-62; SAE 146. (2) Zhao, F.; Asumus, T.; Assanis, D.; Dec, J.; Eng, J.; Najt, P. Homogeneous Charge Compression Ignition (HCCI) Engines : Key Research and DeVelopment Issues; PT-94, SAE: Warrendale, PA, 2003. (3) Yeom, K.; Jang, J.; Bae, C. Homogeneous Charge Compression Ignition of LPG and Gasoline using Variable Valve Timing in an Engine. Fuel 2007, 86 (4), 494-503. (4) Kaiser, E.; Yang, J.; Culp, T.; Xu, N.; Maricq, M. Homogeneous Charge Compression Ignition Engine-Out Emissions-Does Flame Propagation Occur in Homogeneous Charge Compression Ignition? Int. J. Engine Res. 2002, 3 (4), 185-196.

mixture temperature and pressure. A high combustion pressure and a high rate of pressure increase are due to simultaneous self-ignition. The engine operation is limited by the high combustion pressure and the high rate of pressure increase.1,2,5 Some fuels, such as gasoline, natural gas, and liquefied petroleum gas (LPG), do not readily allow auto-ignition, due to their low cetane number and lack of heat energy during the compression period. In order to ignite a low cetane number fuel, a high cetane number fuel was used for ignition promotion.6,7 Hot residual gas is used as a source of heat and can promote HCCI combustion.8-11 This hot residual gas can be controlled (5) Hiraya, K.; Hasegawa, K.; Urushihara, T.; Iiyama, A.; Itoh, T. A Study on Gasoline Fueled Compression Ignition Engine ∼ A Trial of Operating Region Expansion∼. SAE Trans. 2002, 111 (3), 854-862; SAE 416. (6) Xingcai, L.; Yuchun, H.; Libin, J.; Linlin Z.; Zhen, H. Heat Release Analysis on Combustion and Parametric Study on Emissions of HCCI Engines Fueled with 2-Propanol/n-Heptane Blend Fuels. Energy Fuels 2006, 20, 1870-1878. (7) Kim, D.; Kim, M.; Lee, C. Effect of Premixed Gasoline Fuel on the Combustion Characteristics of Compression Ignition Engine. Energy Fuels 2004, 18, 1213-1219. (8) Allen, J.; Law, D. Variable Valve Actuated Controlled AutoIgnition: Speed Load Maps and Strategic Regimes of Operation. SAE Tech. Pap. Ser. 2002, 2002-01-0422. (9) Wolters, P.; Salber, W.; Geiger, J.; Duwsmann M.; Dilthey, J. Controlled Auto Ignition Combustion Process with an Electromechanical Valve Train. SAE Trans. 2003, 112 (3), 160-168; SAE 0032. (10) Standing, R.; Kalian, N.; Ma, T.; Zhao, H.; Wirth, M.; Schamel, A. Effects of Injection Timing and Valve Timing on CAI Operation in a Multi-Cylinder DI Gasoline Engine. SAE Tech. Pap. Ser. 2005, 2005-010132. (11) Yap, D.; Megaritis, A.; Wyszynski, M. An Investigation into Bioethanol Homogeneous Charge Compression Ignition (HCCI) Engine Operation with Residual Gas Trapping. Energy Fuels 2004, 18, 13151323.

10.1021/ef070076h CCC: $37.00 © 2007 American Chemical Society Published on Web 06/22/2007

Combustion and Knock of Gasoline-DME HCCI Engine Table 1. Engine Specifications bore (mm) stroke (mm) compression ratio displacement (cc) intake/exhaust valve duration intake/exhaust valve lift

82 93.5 12 494 228/228 8.5/8.4

Valve Timing (CAD) intake valve open (ATDC) -29 ∼ 11 intake valve close (ABDC) 59 ∼ 19 exhaust valve open (BBDC) 42 exhaust valve close (ATDC) 6 DME injection pressure (MPa) 5 gasoline injection pressure (MPa) 0.3 DME injector slit injector gasoline injector multihole port fuel injector

by a variable valve timing (VVT) device.12 The VVT device can improve volumetric efficiency by varying the intake valve’s open and close timing.13 Di-methyl ether (DME) is suitable as an auto-ignition promoter, due to its high cetane number, and some researchers consider it as an alternative fuel for diesel. The properties of DME are very similar to diesel. The cetane number of DME is higher than that of diesel. Because of its high cetane number, DME is suitable as an ignition promoter in this research. Other major merits of DME are contained oxygen molecules and easy vaporization.14 The easy evaporation avoids wall wetting at early injection. The contained oxygen molecules oxidize soot during the expansion stroke. In order to meet the wide operating range requirement of fuel, dual-fuel HCCI combustion was introduced. The combination of low-cetane number fuel and high-cetane number fuel can satisfy both high- and low-load engine operating conditions simultaneously. The well-known dual-fuel combinations are gasoline-diesel and liquefied petroleum gas (LPG)-DME.2 Most dual-fuel HCCI combustion research is attempting to reveal the combustion characteristics, exhaust emissions, and adaptation of the dual-fuel according to engine operating conditions. In this study, gasoline HCCI combustion and emission characteristics near the high load limit were investigated. VVT was used to control residual gas, and DME was used as an ignition promoter by direct injection. The effects of valve timing and fuel quantity on the gasoline HCCI combustion were investigated. The relationships between the combustion pressure and the rate of combustion pressure rise and the knock probability were obtained, in order to determine the engine operating range. In addition, the combustion and emission characteristics were investigated. Experiments Research Engine. The specifications of the engine are given in Table 1. The base engine is a four cylinder spark ignition (SI) engine equipped with a double overhead camshaft (DOHC) and a VVT device. A cylinder was modified for HCCI combustion operation with a DME direct injection system on the cylinder head together with a port fuel injection system. An engine control unit (ECU; (12) Babajimopoulos, A.; Assanis D.; Fiveland, S. An Approach for Modeling the Effects of Gas Exchange Processes on HCCI Combustion and Its Application in Evaluating Variable Valve Timing Control Strategies. SAE Trans. 2002, 111 (4), 1794-1809; SAE 2829. (13) Jiang, H.; Wang J.; Shuai, S. Visualization and Performance Analysis of Gasoline Homogeneous Charge Induced Ignition by Diesel. SAE Tech. Pap. Ser. 2005, 2005-01-0136. (14) Yu, J.; Bae, C. Dimethyl Ether (DME) Spray Characteristics in a common-rail Fuel Injection System. Proc. Inst. Mech. Eng. Part D: J. Automobile Eng. 2003, 217 (D12), 1135-1144.

Energy & Fuels, Vol. 21, No. 4, 2007 1943 Table 2. Experimental Conditions engine speed (rpm) intake valve open timing (CAD) DME injection timing (CAD) gasoline injection timing (CAD) λTOTAL λDME λgasoline intake charge temperature (°C) coolant/oil temperature (°C)

1000 -29, -19, -9, 1, 11 110 470 2.12, 2.41, 2.57, 2.77, 2.91 3.70 4.96, 6.91, 8.42, 11.02, 13.63 30 80/80

Motec Co., M4) was employed to precisely control the DME quantity and injection timing. A second ECU (ETAS Co.) was used to control the port injection fuel quantity, timing, and intake valve timing. Figure 1 shows a schematic diagram of the experimental setup. The engine speed and load were controlled by an alternating current (AC) dynamometer. A slit injector (Denso Co.) was used to inject the DME directly into the cylinder at a constant supply pressure of 5 MPa using a pressurized nitrogen gas. The DME injector was located at the spark plug hole of the research engine. A lubricity enhancer (Infineum, R655) of 500 ppm was added to the DME, in order to prevent damage to the fuel injection system.14 The incylinder pressure was measured by a piezoelectric pressure transducer (Kistler, 6052b). The short time drift of the cylinder pressure sensor due to thermal shock is under (0.05 MPa. The piezoelectric pressure transducer for in-cylinder pressure measurement was flash-mounted inside the cylinder head to reduce pipe oscillation effects.15 In order to analyze the in-cylinder data precisely, a high-precision rotary encoder (2048 pulse/rev) was used for engine control and data acquisition. The typical pressure oscillation frequency of knock is approximately 5 kHz. The pressure data were acquired at every rotary encoder pulse, which is approximately 34 kHz at 1000 rpm of engine speed, which can avoid the distortion of pressure data. The intake and exhaust manifold pressures were measured by two piezoresistive pressure transducers (Kistler, 4045A5). The intake and exhaust temperature were measured by two K-type thermocouples, which were fitted on the intake and exhaust manifold. A wide band lambda meter (ETAS, LA4) was installed for measurement of the air excess ratio. Exhaust gases were analyzed with a gas analyzer (Horiba, Mexa 1500d) to measure HC, NOx, CO, and CO2 emissions. The air flow rate was measured using a laminar flow meter (Meriam Co., 50MC2-2S), in order to obtain volumetric efficiency. A data acquisition system (IOtech, Wavebook 512H) was employed to acquire all engine combustion and exhaust gas data. Analog to digital (A/D) conversion resolution, speed, and accuracy are 12 bit, 1 MHz, and (0.025% of the full scale, respectively. The combustion, intake, and exhaust pressure data were acquired with a 34 kHz sampling rate. The other data including the mass air flow rate and exhaust gases were acquired with a 1 kHz sampling rate. The indicated mean effective pressure (IMEP) and heat release rate were calculated from the cylinder pressures.16 The VVT system can freely vary the open and close timing of the intake valve. The intake valve open (IVO) timing was varied for a range of 29 crank angle degrees (CAD) before top dead center (BTDC) to 11 CAD after top dead center (ATDC), while the valve duration was fixed at 228 CAD. Generally, the volumetric efficiency and residual gas are increased at low engine speeds as IVO timing is advanced. Increased volumetric efficiency and residual gas promote the combustion.17 (15) Bertola, A.; Stadler, J.; Walter, T.; Wolfer, P.; Gossweiler C.; Rothe M. Pressure indication during knock conditions. Proceedings of the 7th Internal Symposium on Internal Combustion Diagnostics Kurhaus BadenBaden, Germany, May 18-19, 2006; pp 7-21. (16) Heywood, J. B. Internal Combustion Engine Fundamentals; McGrawHill: New York, 1988. (17) Cavina, N.; Ponti, F.; Siviero, C.; Suglia, R. Residual Gas Fraction Estimation for Model-Based Variable Valve Timing and Spark Advance Control. ASME Int. Comb. Engine DiV. 2004, ICEF2004-0956, 457-466.

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Figure 1. Experimental apparatus.

ATDC. At an IVO timing of -29 CAD ATDC, the volumetric efficiency was 80%. At 11 CAD ATDC for IVO timing, it was 66.2%. The volumetric efficiency was calculated using eq 2.16 ην )

Experimental Conditions. Table 2 shows the main experimental conditions used in this study. Air excess ratio, λ, is defined as the proportion of actual air-fuel ratio to stoichiometric air-fuel ratio, (A/F)actual/(A/F)stoichiometric.16 The air-fuel ratio, A/F, is defined as the ratio m˘ a/m˘ f.16 In order to define the concentration of each fuel (λGasoline,λDME), the total mass of fresh air and the fuel quantity of each fuel were used. The total air excess ratio, λTOTAL, is defined as, eq 1, which was derived from the chemical formula of the combustion between a gasoline-DME mixture and air. λGasolineλDME λGasoline + λDME

(2)

Where, ma is the mass air flow rate, Fa,i is the inlet air density, and Vd is the displacement volume. The mass flow rate of the intake charge was measured using a laminar flow meter (Meriam Co., 50MC2-2S). The DME injection timing was fixed at 110 CAD ATDC. The homogeneity of the DME-air mixture was obtained through early injection before 110 CAD ATDC in the DME injection timing. Gasoline was injected at the intake manifold at 470 CAD ATDC. The indicated mean effective pressure (IMEP) and heat release rate were calculated from the cylinder pressures16 by,

Figure 2. Intake valve open and DME injection timing.

λTOTAL )

ma Fa,iVd

(1)

Where, λTOTAL is the total air excess ratio, λGasoline is the gasoline air excess ratio, and λDME is the DME air excess ratio. The engine was run at 1000 rpm for various intake valve timings and air excess ratios. The operating range of an HCCI engine is limited at high-speed operating conditions because of heat transfer and the rise of combustion pressure.2 The operating range of the HCCI engine in terms of engine load is wider when the engine speed is low.2 For these reasons, this research was performed mainly under low engine speed at 1000 rpm. Figure 2 shows the intake valve timings and the DME injection timing. The intake valve open (IVO) timings were varied from -29 CAD ATDC to 11 CAD

dQ γ dV 1 dP ) P + V + ∆Qheattransfer dθ γ - 1 dθ γ - 1 dθ

(3)

where, θ is the crank angle degree, γ is the specific heat ratio, P is the cylinder combustion pressure, Q is the heat release, V is the cylinder volume, and ∆Qheattransfer is the heat transfer from burned gas to cylinder wall. The HCCI oxidation reaction has a two-stage auto-ignition process: heat release with low temperature reaction (LTR) and high temperature reaction (HTR).18,19 The heat release of LTR was approximately 8-12% of the total heat release. To eliminate the effects of LTR, the burn duration (θ20-90) is defined as the period between the position at which 20% of the mass fraction is burned (MFB) and the position at which 90% of MFB. In order to quantify the knock intensity, the ringing intensity (RI) was introduced.20 The use of RI for HCCI knock intensity has been reported as21,22 (18) Sato S.; Iida, N. Analysis of DME Homogeneous Charge Compression Ignition Combustion. SAE Tech. Pap. Ser. 2003, 2003-01-1825. (19) Lee, K.; Lee, C.; Ryu J.; Kim, H. An Experimental Study on the Two-Stage Combustion Characteristics of a Direct-Injection-Type HCCI Engine. Energy Fuels 2005, 19, 393-402. (20) Eng, J. Characterization of Pressure Waves in HCCI Combustion. SAE Tech. Pap. Ser. 2002, 2002-01-2859. (21) Sjoberg, M.; Dec, J. Effects of Engine Speed, Fueling Rate, and Combustion Phasing on the Thermal Stratification Required to Limit HCCI Knock Intensity. SAE Trans. 2005, 114 (3), 1472-1486; SAE 2125.

Combustion and Knock of Gasoline-DME HCCI Engine RI )

1 xγRT 1 2γ P

N

[ ()]

∑ 0.05 N 1

∂p ∂t

Energy & Fuels, Vol. 21, No. 4, 2007 1945

2

(4)

max

where, γ is the specific heat ratio, R is the gas constant, T is the temperature, N is the number of cycles, and (∂p/∂t)max is the maximum pressure rise. The knock probability was calculated using eq 5. Knock probability (%) )

number of cycles with knock × 100 (5) 100 cycles

Figure 3 shows the knocking criterion that was used in this study. The knock intensity is obtained from the cylinder pressure data using various methods.23 The present study focuses on the knock probability. In order to identify the knock, the pressure oscillation criterion due to the knock was determined as the pressure rise over 0.05 MPa during the expansion stroke following the work of Oakley et al.23 The combustion pressures were analyzed for 100 cycles at each experimental condition, in order to obtain the IMEP and knock probabilities.

Figure 3. Criterion of knocking detection in this study. (a) Effects of λTOTAL at a fixed IVO timing (-29 CAD ATDC). (b) Effects of IVO timing at a fixed λTOTAL (2.12).

Results and Discussion Combustion Characteristics of the Gasoline-DME HCCI Engine. The cylinder combustion pressure and the heat release rate, with respect to IVO timing and mixture strength, are shown in Figure 4. The horizontal axis represents the crank angle degree (CAD), and the vertical axis represents the cylinder combustion pressure and the heat release rate. The λTOTAL was varied from 2.91 to 2.12 under a fixed IVO timing of -29 CAD ATDC, and the decrease in λTOTAL yielded an increase in the peak combustion pressure and rate of pressure rise (Figure 4a). However, the change in the start of the combustion for various λTOTAL values between 2.12 and 2.91 was negligible. This is because the DME injection quantity, volumetric efficiency, and residual gas, all of which are related to the start of the combustion process, were not changed. However, when a larger amount of gasoline was injected, the combustion became more vigorous with the higher combustion pressure. The IVO timing was varied from -29 to 11 CAD under a fixed λTOTAL of 2.12. An increased peak combustion pressure and advanced position of peak pressure were observed with advanced IVO timing, due to a higher volumetric efficiency and a larger amount of residual gas (Figure 4b). The amount of residual gas increased as the IVO timing advanced, due to increased duration of valve overlap; the increased residual gas can supply heat to promote LTR.24 The residual gas increased as the valve overlap period was increased20. The valve overlap period is the duration in which the intake valve and the exhaust valve are open simultaneously. The burned gas is trapped inside the cylinder by the pressure difference between the intake port and the exhaust port. The heat transfer from hot residual gas to the fresh charge increased, which promotes ignition.23 The charge temperature reaches a low temperature oxidation temperature, which is 600-900 K1, more rapidly when the residual gas is trapped more in the combustion chamber. The volumetric efficiency increased as the IVO timing was advanced. The (22) Helmantel A.; Denbratt, I. HCCI Operation of a Passenger Car DI Diesel Engine with an Adjustable Valve Train. SAE Tech. Pap. Ser. 2006, 2006-01-0029. (23) Oakley, A.; Zhao, H.; Ladommatos, N.; Ma, T. Experimental Studies on Controlled Auto-Ignition (CAI) Combustion of Gasoline in a 4-Stroke Engine. SAE Trans. 2001, 110 (3), 1062-1075; SAE 1030. (24) Caton, P. A.; Simon, A. J.; Gerdes, J. C.; Edwards, C. F. Residualeffected Homogeneous Charge Compression Ignition at a Low Compression Ratio using Exhaust Reinduction. Int. J. Engine Res. 2003, 4 (3), 163178.

Figure 4. Cylinder combustion pressure and heat release rate of gasoline and DME HCCI engine at 1000 rpm. (a) Effects of λTOTAL at a fixed IVO timing (-29 CAD ATDC). (b) Effects of IVO timing at a fixed λTOTAL (2.12).

volumetric efficiency was 80%, 79.5%, and 77.2% at -29, -19, and -9 CAD ATDC, respectively. However, the volumetric efficiency dropped rapidly at 1 and 11 CAD ATDC to 70.8% and 66.2%, respectively. The fresh charge temperature reached 800-1000 K easily at the compression stroke due to the increased volumetric efficiency. The increased volumetric efficiency and residual gas amounts are attributed to early ignition when the IVO timing was retarded. The HCCI combustion starts from a low temperature reaction (LTR), which supplies heat and pressure to the air-fuel mixture. This causes the gasoline to reach the self-ignition point.17 Figure 5 shows the heat release rate of the LTR before main combustion occurs. Under fixed IVO timing, the LTR heat release decreased

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Figure 6. IMEP of the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm.

Figure 5. Heat release rate of LTR of the gasoline and DME HCCI engine at 1000 rpm. (a) Effects of λTOTAL at a fixed IVO timing (-29 CAD ATDC). (b) Effects of IVO timing at a fixed λTOTAL (2.12).

and was retarded as λTOTAL was decreased (Figure 5a). The LTR timing is slightly affected by the gasoline injection quantity. The mixture temperature is lowered by the latent heat of vaporization of gasoline during the compression stroke. This temperature decrease is a dominant factor in the LTR retardation.2 Under a fixed λTOTAL, the heat release rate of the LTR decreased as the IVO timing was retarded (Figure 5b). The retardation is attributed to the reduction in residual gas and volumetric efficiency. The indicated mean effective pressures (IMEPs), with respect to the λTOTAL and the IVO timing, are shown in Figure 6. The horizontal axis represents the IVO timing and the vertical axis represents the λTOTAL. This figure shows that λTOTAL is the major influencing factor for the IMEP. There is an optimum amount of injected gasoline for attaining optimal IMEP, corresponding to an approximate value of 2.4 in λTOTAL. When λTOTAL was lower than 2.4, early combustion was observed and IMEP dropped rapidly. This is because most of the heat was released before TDC and the negative work during compression was increased as λTOTAL was decreased. IMEP is increased as the IVO timing is retarded when the λTOTAL was lower than 2.4. The volumetric efficiency and residual gas decrease, due to late intake valve opening and closing. This reduced volumetric efficiency and residual gas lead to late combustion, which reduces the negative work. Figure 7 shows the θ20-90 of the gasoline HCCI engine, which was mainly influenced by λTOTAL rather than by the IVO timing. This is because the injected fuel energy is a critical factor in θ20-90. However, in the case of IVO timing at 11 CAD ATDC, θ20-90 was slightly longer than that of the early IVO timing

Figure 7. Burn duration of the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm.

Figure 8. Ringing intensity of the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm.

conditions, due to the lowered volumetric efficiency and the reduction of residual gas. Knocking Characteristics of the Gasoline-DME HCCI Engine. Figure 8 shows the ringing intensity (RI) of the gasoline and DME HCCI engine with respect to λTOTAL and intake valve open (IVO) timing at 1000 rpm. The horizontal axis represents the IVO timing, and the vertical axis represents λTOTAL. The RI

Combustion and Knock of Gasoline-DME HCCI Engine

Figure 9. Knock probability of the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm.

increased as the IVO timing was advanced. This is attributed to increased volumetric efficiency and residual gas which promote combustion. More than 50% of the fresh charge was burned, and the combustion pressure was rising before TDC at an early IVO timing. The fresh charge temperature reached 800-1000 K easily due to the increased volumetric efficiency at the compression stroke. The residual gas increased as the valve overlap period was increased.17 The allowable operating range based on audible knock sounds is regarded as being under 5 MW/m2 of RI which is defined as the threshold value of the high load operating range limit.21,22 In the case of gasoline HCCI, the RI was over 5 MW/m2, when λTOTAL was lower than 2.3, and the IVO timing was more advanced than -10 crank angle degrees. The RI, based on the indicated mean effective pressure (IMEP), was more important for better fuel economy than that based on the audible knock sounds. Figure 6 shows the IMEP with respect to λTOTAL and the IVO timing at 1000 rpm. The relationship between the IMEP and the RI showed that the IMEP of the gasoline HCCI engines decreased by 10% when the RI was over 0.5 MW/m2. According to Figures 6 and 8, it was found that the IMEP was significantly affected by the RI value. The IMEP of the gasoline-DME engine was decreased when the RI value was over 0.5 MW/m2. This fact indicates that the high engine load should be limited to a certain range where the RI value is less than 0.5 MW/m2. Figure 9 shows the knock probability of the gasoline HCCI. Through 100 cycles, the cylinder pressure data for each experimental condition were used for the knock probability analysis. The knock probability increased as the λTOTAL decreased and the IVO timing advanced. The reduced λTOTAL and advanced IVO timing lead to a higher combustion pressure and an increased rate of combustion pressure rise. Figure 10 shows the maximum combustion pressure (Pmax), with respect to the engine operating conditions. The horizontal axis represents the IVO timing and the vertical axis represents the λTOTAL. Figure 10 shows that Pmax increases as λTOTAL is decreased and the IVO timing is advanced. Figure 11 presents the maximum pressure rise in combustion, with respect to the λTOTAL and IVO timing. The maximum pressure rise during combustion increased with the decreased λTOTAL and the advanced IVO timing. The knock probability of the HCCI combustion also increased as the maximum combustion pressure rise increased, due to the vigorous combustion resulting from the rich mixture. Moreover, early intake

Energy & Fuels, Vol. 21, No. 4, 2007 1947

Figure 10. Maximum combustion pressure of the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm.

Figure 11. Maximum pressure rise of the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm.

Figure 12. Knock probability of the gasoline and DME HCCI engine with respect to maximum combustion pressure at 1000 rpm.

valve closing allows more charge into the cylinder and traps more hot residual gas in the cylinder. Hence, Pmax and a maximum combustion pressure rise are key parameters in the knocking probability of the HCCI engine combustion. The knock probability, with respect to Pmax, is shown in Figure 12. Figure 13 shows the knock probability, with respect to the maximum combustion pressure rise. The relation of the knock probability (y) and the Pmax (x) can be fit by the linear eq 6,

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Figure 13. Knock probability of the gasoline and DME HCCI engine with respect to maximum combustion pressure rise at 1000 rpm. Figure 15. CO2 emission from the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm.

Figure 14. HC emission from the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm. Table 3. Polynomial Fit Constants

Figure 16. CO emission from the gasoline and DME HCCI engine with respect to λTOTAL and IVO timing at 1000 rpm.

constants A B1 B2 B3 B4

-93.03 1811.06 -5627.89 3934.15 5666.98

through 25 engine operating conditions, which were tested in this study.

y ) 47.20x - 215.15

(6)

where, x is the maximum combustion pressure (Pmax) and y is the knock probability. The relation of the knock probability (y) and the maximum combustion pressure rise (z) can be also fit by eq 7. The constants of eq 7 are shown in Table 3.

y ) A + B1z + B2z2 + B3z3 + B4z4

(7)

where, y is the knock probability and z is the maximum combustion pressure rise, (dP/dθ)max. Exhaust Emissions of Gasoline-DME HCCI Engine. The most important advantages of an HCCI engine in terms of exhaust emission are much less particulate matter (PM) and nitrogen oxide (NOx) emissions resulting from homogeneous charge ultralean mixtures.2 Figure 14 shows the HC emission from a gasoline HCCI engine, with respect to the λTOTAL and the IVO timing. The HC emission increased as the IVO timing

was retarded. The increased HC emission was due to delayed combustion. Late combustion leads to a lower combustion pressure and temperature. The fuel that was not burned during the combustion process becomes the source of the HC emission. The HC emission could be oxidized by the heat from the burned gas during the expansion stroke. However, the HCCI engine combustion temperature is lower than that of the SI and CI engines. Thus, the main causes of increased HC emission from the HCCI engine are the strong quenching effect and the lack of an oxidation reaction during the expansion stroke. In addition, in the case of late IVO timing, the quenching effect was stronger than that for early IVO timing, due to a low combustion temperature. This implies that a lower combustion temperature creates higher HC emission, through reduced oxidation and increased quenching, in the case of late IVO timing relative to early IVO timing. Figure 15 presents a map of the CO2 emission. The CO2 emission was influenced mainly by λTOTAL. This is because of the fuel quantity. At the upper right region of the CO2 emission map, CO2 emission decreased as the IVO timing was retarded. This is related to the CO dissociation. In order to illustrate the relation of the CO and CO2 emissions, a CO emission map is shown in Figure 16. The upper right region of the CO emission map shows a higher CO concentration than the other regions of the map. The

Combustion and Knock of Gasoline-DME HCCI Engine

combustion temperature was lower than the early IVO timing and lower λTOTAL conditions. The oxidation reaction during the expansion stroke was also weaker than the other conditions. However, in the case of HC emission, the HC emission increased as λTOTAL was decreased. This is because the increasing rate of HC emission produced during the combustion process was larger than the increasing rate of oxidized HC emission during the expansion stroke. The NOx emission from the test engine was at almost a zero level, less than 2 ppm, along all the experimental conditions. The PM emission was also negligible at less than 0.01 filter smoke number (FSN) along all the experimental conditions. Conclusions The effects of intake valve open (IVO) timing and total air excess ratio (λTOTAL) on exhaust emissions and combustion characteristics in a gasoline-fueled homogeneous charge compression ignition (HCCI) engine controlled by variable valve timing (VVT) were investigated to find the high-load limits in HCCI operation. A heat release analysis was performed, in order to verify the HCCI combustion characteristics. The knock characteristics were also investigated. The relation between the knock probability and the combustion pressure was determined, in order to obtain the operating parameters for the high load limit. The following conclusions were drawn from these experiments. (1) The start of combustion is independent of the gasoline injection quantity. This is because the di-methyl ether (DME) injection quantity is initially the dominant factor for combustion. However, the burn duration (θ20-90) and the rate of combustion pressure rise are dependent on the fuel quantity and the IVO timing. (2) The indicated mean effective pressure (IMEP) was determined by the fuel quantity. There is an optimal fuel quantity for the best IMEP, corresponding to an approximate value of 2.4 for the total air excess ratio (λTOTAL). The best IMEP is approximately 0.28 MPa at 1000 rpm. (3) The knocking probability and intensity increased as the peak combustion pressure and the combustion pressure rise increased. According to the IMEP decrease when λTOTAL was lower than 2.4, the high load operating range of the test engine was limited to under 0.5 MW/m2. The knock probability can be predicted by two derived equations. (4) The hydrocarbon (HC) emission increased as the IVO timing was retarded and λTOTAL decreased. The main causes of increased HC emission from the HCCI engine are the strong quenching effect and the lack of an oxidation reaction during the expansion stroke. A decreased combustion temperature created more HC emission in the case of late IVO timing relative to that of early IVO timing. (5) The carbon dioxide (CO2) emission is influenced mainly by λTOTAL, which is directly related to the fuel quantity. The CO2 emission from an HCCI engine is basically lower than that of conventional spark ignition (SI) and compression ignition

Energy & Fuels, Vol. 21, No. 4, 2007 1949

(CI) engines due to the leaner combustion and less oxidation of carbon monoxide (CO) during the expansion stroke. The high CO emission from HCCI combustion is one of the main topics of HCCI combustion for further research, which is needed to clarify the major reduction mechanism of CO2 emission. Acknowledgment. The authors would like to show appreciation to the Combustion Engineering Research Center (CERC) at KAIST and engine research center in the Korea Institute of Machinery and Materials (KIMM) for the financial support.

Nomenclature λTOTAL ) total air excess ratio λGasoline ) gasoline air excess ratio λDME ) DME air excess ratio ma ) mass air flow rate Fa,i ) inlet air density Vd ) displacement volume θ ) crank angle degree P ) cylinder combustion pressure Q ) heat release V ) cylinder volume ∆Qheattransfer ) heat transfer from burned gas to cylinder wall θ20-90 ) burn duration γ ) specific heat ratio R ) gas constant T ) temperature N ) number of cycles (∂p/∂t)max ) maximum pressure rise x ) maximum combustion pressure (Pmax) y ) knock probability z ) maximum combustion pressure rise (dP/dθ)max AbbreViations AC ) alternating current ABDC ) after bottom dead center ATDC ) after top dead center BBDC ) before bottom dead center BTDC ) before top dead center CAD ) crank angle degree CI ) compression ignition CO ) carbon oxide CO2 ) carbon dioxide DI ) direct injection DOHC ) double over head camshaft DME ) di-methyl ether ECU ) engine control unit HC ) hydrocarbon HCCI ) homogeneous charge compression ignition IMEP ) indicated mean effective pressure IVO ) intake valve open LHV ) low heating value MFB ) mass fraction burned NOx ) nitrogen oxides SI ) spark ignition VVT ) variable valve timing EF070076H