Combustion Characterization in an Internal Combustion Engine with

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VOLUME 23

MAY 2009 Copyright 2009 by the American Chemical Society

ReViews Combustion Characterization in an Internal Combustion Engine with Ethanol-Gasoline Blended Fuels Varying Compression Ratios and Ignition Timing Christopher P. Cooney,* Yeliana , Jeremy J. Worm and Jeffrey D. Naber Michigan Technological University, 1400 Townsend Drive, Houghton, Michigan 49931 ReceiVed October 23, 2008. ReVised Manuscript ReceiVed March 3, 2009

Although ethanol possesses only two-thirds the energy density of gasoline, it has other properties that are beneficial to combustion in an internal combustion (IC) engine. These include a higher laminar flame speed and higher octane number relative to gasoline. The higher octane number of ethanol improves knock tolerance, and the faster flame speed provides potential benefits to the combustion process. Understanding these attributes will enable flexible fuel engines to benefit from some of the unique properties of ethanol. While data concerning the efficiency of ethanol fuels at varying compression ratios exist in the literature, there is a lack of fundamental combustion data to validate these results. As such, this work makes use of mass fraction burn (MFB) analysis to examine the differences between various blends of ethanol and gasoline, including the effect of the compression ratio relative to optimal combustion phasing, early and bulk burn rates, and combustion variability. Tests were carried out on a port-fuel-injected, electronically-controlled, modified single-cylinder cooperative fuels research engine operating at steady-state conditions and a stoichiometric air/fuel ratio. Combustion experiments were conducted at a constant engine load of 330 kPa net indicated mean effective pressure (NIMEP). To characterize the combustion process, in-cylinder pressure data were used to calculate MFB profiles. Combustion was examined as a function of the ethanol concentration, spark timing, and compression ratio. The experimental results indicated that higher ethanol blends increased the knock-limited compression ratio (KLCR). KLCR for the speed and load tested is 8 for gasoline with an octane rating of 91 research octane number (RON) as compared to 16 for an ethanol blend of 84% ethanol (E84). With an increased ethanol concentration in the fuel, the 0-10% MFB duration decreased. Of all of the fuel blends tested, pure gasoline had the longest burn duration in both 0-10% and 10-90% MFB intervals.

Introduction The potential of increased availability of ethanol as a fuel from renewable sources has major automobile manufacturers developing flex-fuel vehicles designed to run on gasoline and E85 (a mixture of 85% ethanol and 15% gasoline).1 One of the more attractive properties of ethanol as a fuel is that it can be produced from renewable sources, such as agricultural feedstock and cellulosic biomass, or any material that can be transformed into fermentable sugar.2 Ethanol has physical and chemical properties that differ from gasoline. The lower heating value of ethanol on a volume basis * To whom correspondence should be addressed. E-mail: [email protected]. (1) Wicker, R. B.; Hutchison, P. A.; Acosta, O. A.; Matthews, R. D. Practical considerations for an E85-fueled vehicle conversion. SAE Tech. Pap. 1999-01-3517, 1999. (2) Thring, R. H. Alternative fuels for spark-ignition engines. SAE Tech. Pap. 831685, 1993.

is approximately 65% that of gasoline. This leads to decreased fuel economy in automobiles.1 This is also evident in the differences of stoichiometric air/fuel ratios, where ethanol is 9.0:1 and gasoline is 14.5:1. However, for use in an internal combustion (IC) engine, the increase in fuel consumption cannot be predicted by looking at the lower heating value (LHV) alone. Ethanol has other properties that make it an attractive fuel for use in MIC engines. As shown in Table 1, ethanol has a higher octane number than gasoline. This difference allows for engine operation at a greater compression ratio, resulting in increased thermal efficiency without an increased hazard of knock.1,3 Also, ethanol has been measured to have a higher laminar flame speed, which could provide benefits to combustion in cases of increased residual mass fraction and increased levels of exhaust gas recirculation (EGR). A study by Farrell et al. provided the calculated laminar burning of 45 different hydrocarbons at various equivalence ratios at elevated temperature and pressure of T ) 450 K and

10.1021/ef800899r CCC: $40.75  2009 American Chemical Society Published on Web 04/02/2009

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Energy & Fuels, Vol. 23, 2009 Table 1. Properties of Gasoline and Ethanola property

gasoline

ethanol

chemical formula octane number (RON) octane number (MON) latent heat of vaporization (kJ/kg) lower heating value (MJ/kg) stoichiometric A/F laminar flame speed (m/s)b

C4-C12 91-99 82-89 350 44 14.6 0.333

C2H5OH 107 89 840 26.9 9.0 0.388

a From ref 12 and Hara, T.; Tanoue, K. Laminar flame speeds of ethanol, n-heptane, iso-octane air mixtures. Technical Report, International Federation of Automotive Engineering Societies (FISITA), F2006SC40. b Equivalence ratio of 1.0, initial pressure of 100 kPa, and temperature of 325 K. The gasoline value is assumed from iso-octane data.

P ) 304 kPa.4 Data, at an equivalence ratio of 1, show isooctane with a laminar burning velocity of 0.58 m/s and ethanol with a laminar burning velocity of 0.85 m/s. Data of laminar flame speed at lower temperature and pressure are shown in Table 1. A theoretical study by Bayraktar reported that ethanol addition to gasoline up to 25% volume accelerated the flame propagation process.5 Key properties of ethanol are also shown in Table 1. Mass Fraction Burn (MFB) Analysis In this work, combustion was examined via the MFB profile. The MFB profile represents the cumulative percentage of fuel consumed during the combustion process and is normally plotted as a measure of crank slider position. It has a characteristic S shape beginning at the spark event, then increasing to a maximum gradient approximately halfway through the combustion process, and gradually ending as the flame front extinguishes. The profile of MFB is useful in that it provides an indication of the rate at which energy is released during combustion and is influenced by parameters such as the engine design, operating conditions, control variables, and fuel type. The MFB curve is commonly used to study and characterize the effects of dilution on combustion. For example, in a study by Ayala et al., MFB data were used to examine the effects of dilution and hydrogen enhancement on engine efficiency.6 The effect of γ (the ratio of specific heat capacity of gases) in the MFB calculation has been an interesting topic over the years. Krieger et al. used a γ constant along the combustion process. Gatowski et al. investigated a linear correlation of γ with the temperature. Then, Chun and Heywood computed the γ from the average γ of the two-zone model.7-9 With the correct γ, the singlezone heat release model is robust.10 On the basis of the conclusions of Cheung and Heywood, it was decided to use a single-zone model for MFB analysis in this work.10 In the single-zone model used, γ is calculated as a function of the burned gas temperature. Furthermore, the model includes the effects of heat transfer and crevice volume. Additional details on the MFB model used in this study can be found in Yeliana et al.11 MFB data (3) Brusstar, M.; Stuhldreher, M.; Swain, D.; Pidgeon, W. High efficiency and low emissions from a port-injected engine with neat alcohol fuels. SAE Tech. Pap. 2002-01-2743, 2002. (4) Farrell, J. T.; Johnston, R. J.; Androulakis, I. P. Molecular structure effects on laminar burning velocities at elevated temperature and pressure. SAE Tech. Pap. 2004-01-2936, 2004. (5) Bayraktar, H. Theoretical investigation of flame propagation process in an SI engine running on gasoline-ethanol blends. Renewable Energy 2006, 32, 758–771. (6) Ayala, F. A.; Gerty, M. D.; Heywood, J. B. Effects of combustion phasing, relative air fuel ratio, compression ratio, and load on SI engine efficiency. SAE Tech. Pap. 2006-01-0229, 2006. (7) Krieger, R. B.; Borman, G. L. The computation of apparent heat release for internal combustion engines. ASME Tech. Pap. 66-/WA/DGP4, 1996. (8) Gatowski, J. A.; Balles, E. N.; Chun, K. M.; Nelson, F. E.; Ekchian, J. A.; Heywood, J. B. Heat release analysis of engine data. SAE Tech. Pap. 841359, 1984. (9) Chun, K. M.; Heywood, J. B. Estimating heat-release and mass-ofmixture burned from spark-ignition engine pressure data. Combust. Sci. Technol. 1987, 54, 133–143.

Table 2. CFR Engine Specifications13 compression ratio bore (cm) stroke (cm) connecting rod length (cm) displacement (cm3) IVO IVC EVO EVC maximum speed (rpm)

5.4-18.5 8.26 11.43 25.40 611.2 10° ATDC 34° ABDC 40° BBDC 15° ATDC 900

shown in this paper is an average of the MFB data calculated individually from 300 cycles. The MFB (the S-shaped profile) is divided into three separate phases of combustion, as given by Heywood.12 A brief summary of that description follows: (i) For 0-10% MFB, the early flame development period covers the early stage, which starts from the spark discharge and ends where 10% of the charge has burned. This early burn period reflects the preparation stage of combustion. (ii) For 10-90% MFB, the rapid burning period represents the bulk of the combustion process. In this stage, the major portion of the charge burns at the beginning and end of the early flame development. Combustion propagates across the cylinder to the combustion chamber walls, creating a pressure rise in the cylinder above what would be in the absence of combustion. This is the most important period, which reflects the total useful energy that is generated in the cylinder. (iii) For 90-100% MFB, the flame termination period shows the stage where the flame extinguishes on the combustion chamber surfaces. This period reflects the heat-transfer losses. The location of 50% MFB is the crank angle position in the cycle when 50% of the air/fuel charge is burned. When the MFB profile takes on a characteristic S-shaped curve, this point of reference can fall near the center of the MFB profile, where heat release occurs at its maximum level. Throughout this paper, the author uses this position in the combustion event to reference the phase of combustion relative to top dead center (TDC).

Experimental Setup Test Engine. The experiments were conducted using a cooperative fuels research (CFR) engine manufactured by Waukesha Motor Company. The CFR engine is a four-stroke, single-cylinder, spark-ignition engine, with variable-compression ratio capabilities. The engine compression ratio is quickly and accurately changed by moving the entire cylinder sleeve and head with respect to the piston via a hand crank without affecting valve clearances or the basic combustion chamber configuration. The CFR engine compression ratio adjustment from the manufacturer allows the compression ratio of the engine to be changed within the limits of 4-10. For this work, the ability to operate at higher compression ratios was desired, and the CFR engine was modified by replacing the piston with one at a taller compression height and changing compression ratio limits of the engine between the ranges of 5.4-18.5. A direct-current (DC) dynamometer was used to hold engine speed at a constant 900 rpm. Table 2 shows the specifications of the CFR engine used in the experiments. A Mototron target-based rapid-prototyping (TB-RP) system using modern electronic control hardware was installed on the engine.14 Figure 1 illustrates the sensors and actuators installed (10) Cheung, H. M.; Heywood, J. B. Evaluation of a one-zone burnrate analysis procedure using production SI engine pressure data. SAE Tech. Pap. 932749, 1993. (11) Yeliana; Cooney, C.; Worm, J.; Naber, J. D. The calculation of mass fraction burn rates of ethanol gasoline blended fuels using single and two-zone models. SAE Tech. Pap. 2008-01-0320, 2008. (12) Heywood, J. B. Internal Combustion Engine Fundamentals; McGraw-Hill: New York, 1998. (13) American Society for Testing and Materials (ASTM). Rating motor fuels by motor and research methods. ASTM Manual, 1956. (14) Naber, J. D.; Bradley, E. K.; Szpytman, J. E. Target based rapid prototyping control system for engine research. SAE Tech. Pap. 2006-010860, 2006.

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Figure 2. CFR combustion chamber showing the locations of the spark plug and pressure transducer.

Figure 1. CFR engine diagram, including sensors and actuators.

in the system. The system operates on similar principles as other rapid prototyping software control systems, except that the target is a production engine control module (ECU). The system includes automatic code generation and real-time calibration, monitoring, and logging capability in Mototune. Engine control software is written in Matlab’s Simulink Stateflow work environment. Additional custom toolboxes in Simulink enable connection to the ECU I/O through interfaces to the low-level ECU software. The spark plug was relocated to the top of the cylinder in the location of the original magnetic pick-up plug position. The crank-angle position was measured using a 1024 pulse per revolution optical encoder. Air flow was measured using a laminar flow element (LFE). In addition, an automotive mass air flow (MAF) sensor used for the electronic control was calibrated to the induction system used in the experimental setup. A plenum that has a volume 25 times the displaced volume was installed on the intake manifold to damp the pressure oscillation. A second (375 L) tank was installed downstream of the MAF and LFE to ensure steady airflow through these measurement devices. The oxygen content was measured in exhaust using a NTK universal exhaust gas oxygen (UEGO) sensor. Electronic actuators included a digital ignition coil, port fuel injectors, and an electric throttle valve. The intake manifold pressure was measured using a 0-103 kPa absolute pressure sensor. In-cylinder pressure was measured using an AVL piezoelectric GH 12D pressure transducer equipped with a PH01 flame arrestor and a PCB Model 482C series charge amplifier. The transducer was mounted in an adapter installed in the original spark plug position, located at the top of the cylinder wall, with a 2.0 mm diameter × 3.5 mm long passage between the transducer and the combustion chamber. The change of the spark plug location and the use of a mini-transducer were necessary for high compression ratio tests to avoid hindrance caused by the engine modification from installation of a piston with a taller compression height. It was also necessary to tap the transducer at an angle to create enough clearance for the fit of the socket of the transducer and to make the passageway from combustion chamber to the transducer as short as possible. A scale drawing

of the combustion chamber geometry at compression ratios of 8 and 16 along with the mounting of the pressure transducer and spark plug are shown in Figure 2. The cylinder pressure was referenced to the intake manifold pressure at BDC of the intake stroke. Data Acquisition. A National Instruments PCI 6251 was used to acquire and log in-cylinder pressure, crank encoder A pulse (1024 pulses/revolution), crank encoder Z pulse (1 pulse/ revolution), intake manifold absolute pressure, and spark signal. Data for post-processing was logged at 100 kHz sampling frequency. A low-speed National Instruments NI SCXI-1000DC was used for all other measurements, including temperature and airflow from the laminar flow element. The Mototune calibration tool was used for adjustment and calculation of engine control parameters, including the mass air flow rate, spark timing, injection duration, and equivalence ratio.14 To maintain a constant 330 kPa NIMEP during testing, incylinder pressure and manifold absolute pressure data were acquired using the engine optical encoder as an external clock. Engine IMEP was calculated within data acquisition software for real-time feedback of the engine load. Test Conditions The experiments were conducted by sweeping spark timing, ethanol concentration (E0, E20, E40, E60, and E84), and compression ratio (8, 10 12, 14, and 16) at a constant engine load of 330 kPa NMEP and a constant speed of 900 rpm. With a limited top speed of 900 rpm on the CFR engine, a load of 330 kPa was chosen to be an appropriate match. This part load also creates the conditions for an increase in the degree of dilution of residual burned gases in the cylinder charge. All fuel blends were mixed using unleaded test gasoline (UTG) and laboratory-grade E85 from Chevron Phillips Chemical Company LLC. Fuel specification of E85 indicated an ethanol content of 83.7% and is referred within this document as E84. UTG91 was used to represent the performance of 87 pump octane (R + M/2) gasoline. The engine coolant was maintained at 90 ( 2 °C in a closed system using a shell-and-tube heat exchanger. Building water to cool the heat exchanger was controlled with an electronic proportional control valve and PID controller. The

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Figure 3. Operating range of higher ethanol-blended fuels in terms of KLCR.

oil temperature was maintained at 60 ( 4 °C using a heater secured to the base of the engine and a PID temperature controller. Spark timing was swept in 2 crank angle degree (CAD) increments starting from -4 CAD retarded from TDC, advancing until the location of the peak cylinder pressure was 10 CAD after top dead center (ATDC) or until heavy audible knock, whichever occurred first. Points of borderline knock were identified visually from the real-time pressure trace and noted in the test sheet. Real-time NIMEP data was used to maintain a constant load of 330 kPa NIMEP via adjustments to the throttle. Simultaneously, the fuel injector pulse width was adjusted to maintain an exhaust λ of 1.0, indicated by the UEGO sensor, for each fuel at each set point. After steady-state engine conditions were met, 300 consecutive engine cycles were recorded in time base with a sampling frequency of 100 kHz. Results and Discussion Octane Effects. Referring again to Table 2, ethanol has a higher octane rating than gasoline, indicating greater antiknocking characteristics. When ethanol is blended with gasoline, the antiknock characteristic is increased with the ethanol fraction, and evidence of this is revealed in Figure 3. With an increased ethanol fraction, the knock-limited compression ratio (KLCR) at an engine load of 330 NIMEP is extended. KLCR is defined as the maximum CR for which maximum brake torque (MBT) or, for this part throttle condition, optimal combustion phasing could be achieved without encountering knock. Burn Rate Effects. The majority of experimental studies describing the performance of ethanol blends used in IC engines have been devoted to emission characteristics and engine efficiency and studies of the compression ratio.15-17 As a result, there exists a considerable amount of data on the subject. The higher octane rating gives ethanol a potential for increased engine efficiency via an increased compression ratio. Automobile engines are designed to operate well on the lowest octane fuel that is readily available, which is typically 87 PON, in the United States. Because variable compression ratio engines are not in production, the compression ratios of flexible fuel engines are limited for gasoline use and thus do not take advantage of the higher octane ratings of ethanol blends. Significantly less information is available regarding the effect of ethanol on (15) Popuri, S. S. S.; Bata, R. M. A performance study of iso-butanol, methanol and ethanol-gasoline blends using a single cylinder engine. SAE Tech. Pap. 932953, 1993. (16) Guerrieri, D. A.; Caffrey, P. J.; Rao, V. Investigation into the vehicle exhaust emissions of high percentage ethanol blends. SAE Tech. Pap. 950777, 1995. (17) Al-Farayedhi, A. A.; Al-Dawood, A. M.; Gandhidasan, P. Experimental investigation of SI engine performance using oxygenated fuel. J. Eng. Gas Turbines Power 2004, 126, 178–191.

Figure 4. Comparison of the MFB profile of different ethanol blends at a compression ratio of 8:1, NIMEP of 330 kPa, spark timing of 0 CAD.

combustion rates. In addition, more fundamental data regarding the effect of ethanol concentration on combustion will enable improvements in accuracy of combustion simulation models and 1D engine codes used for flex-fuel engine design. The purpose of this study is to determine how the level of ethanol content in ethanol/gasoline blends effects the MFB profile. The effects of the compression ratio are also shown. It is important to note that turbulence has a strong effect on the combustion process. Details on motored engine turbulence in the CFR combustion chamber can be found in ref 18. Additional work to investigate the relationship between mixture turbulence and combustion rate in the CFR combustion chamber was shown in a corollary work.19 These studies report that turbulence increases flame speed by increasing the surface area of the flame through distortion by wrinkling. Also, turbulence was found to be influenced by engine speed and volumetric efficiency, with levels of average mean velocity and turbulence intensity decreasing with decreased speed and volumetric efficiency. This work also reported that turbulence tended toward isotropy during compression. This work maintains constant speed with very minimal changes in volumetric efficiency. It is expected that fuels with higher flame speed will affect the slope of the various regions of the MFB profile of ethanol gasoline blends. Figure 4 shows a comparison of the MFB profiles with different ethanol concentrations. The curves are normalized to the maximum burned fraction to assist in making the comparison between different fuel blends. The spark timing is the same for all fuels and is at TDC. The engine load is 330 kPa NIMEP at 900 rpm. When the differences in the MFB profiles are examined, it can be seen that the ethanol fraction in the fuel affects the shape of the MFB profile. Gasoline has the longest duration of total MFB (0-100% MFB), while E84 has the shortest. Ethanol blends have an increased gradient with increased ethanol fraction, showing a faster heat release rate. Separation between the profiles is largest between E20 and UTG91, while there is less separation between ethanol blends (E20-E84). Figure 5 shows the difference in MFB profiles with varying compression ratios with E84 as the fuel. Spark timing was again at top dead center (TDC) for all cases and was required in this comparison to avoid knock at a CR of 16. The figure shows that an increased compression ratio creates an increase in the (18) Lancaster, D. R. Effects of engine variables on turbulence in a sparkignition engine. SAE Tech. Pap. 760159, 1976. (19) Lancaster, D. R.; Krieger, R. B.; Sorenson, S. C.; Hull, W. L. Effects of turbulence on spark-ignition engine combustion. SAE Tech. Pap. 760160, 1976. (20) Chevron Phillips UTG-91, Certificate of Analysis. (21) Chevron Philips E-85 Fuel, Certificate of Analysis.

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Figure 5. Comparison of the MFB profile of E84 with varying compression ratios, NIMEP of 330 kPa, and spark advance of 0 CAD.

Figure 6. Comparison of 0-10% MFB durations of spark swept data showing different ethanol blends at a compression ratio of 8:1 and NIMEP of 330 kPa.

gradient in the MFB profile during the early stages of combusiton. The graph shows that, after the initial increase in the burn rate, the rate of heat release slows from the poor combustion chamber geometry as a result of an increased compression ratio. Refer to Figure 2 showing details of the combustion chamber geomety at compression ratios of 8 and 16. Figure 6 shows the 0-10% MFB duration as a function of spark timing. As shown in the figure, initially when the combustion phasing is retarded, the 0-10% MFB duration decreases for all fuels. This is a result of the increased compression during this phase of the burn. As the 0-10% MFB event occurs nearer TDC, it occurs under higher temperature and pressure conditions because of the increased compression. However, as the combustion phasing continues to be retarded and the 0-10% MFB event starts to occur after TDC, the opposite situation occurs and the burn rate slows as the temperature and pressure decrease with the expanding cylinder volume. Figure 7 shows the 10-90% MFB duration in relation to the location of 50% MFB for spark sweep data at a compression ratio of 8:1. Shown on the plot is a near linear relationship (with the exception of very advanced combustion phasing, which was only attainable with E60 and E84) between CA50 and 10-90% MFB. As the combustion phasing is retarded (moving from left to right along the x axis), the 10-90% MFB event occurs at lower pressure and temperature because of the expanding cylinder volume, resulting in slower burn rates. Figures 8 and 9 summarize the test data by plotting the 0-10 and 10-90% MFB durations, respectively, for both the ethanol concentration and compression ratio.

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Figure 7. Comparison of 10-90% MFB durations of spark swept data showing different ethanol blends at a compression ratio of 8:1 and NIMEP of 330 kPa.

Figure 8. Comparison of 0-10% MFB duration as a function of the ethanol concentration and compression ratio, with NIMEP of 330 kPa and MBT combustion phasing.

Figure 9. Comparison of 10-90% MFB duration as a function of the ethanol concentration and compression ratio, with NIMEP of 330 kPa and MBT combustion phasing.

It is also shown that, for any given ethanol concentration, the 0-10% burn duration becomes shorter as the compression ratio is increased. This can be attributed to the higher temperatures at the point of ignition from the increased compression of the fuel air mixture during the compression stroke. In Figure 8, it is shown that the 0-10% MFB tends to decrease as the ethanol concentration increases for the lower compression ratios, while at the higher compression ratios, it is relatively insensitive to the ethanol concentration. The early burn period that is represented by the 0-10% MFB duration can be assumed to be a laminar process, and as such, the increased

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Figure 10. Comparison of MBT combustion phasing as a function of the ethanol fraction for different compression ratios, with NIMEP of 330 kPa.

laminar flame speed of ethanol is shown to have a small effect, albeit at low compression ratios. However, at the higher compression ratios, the effect of the compression work and, therefore, charge density and temperature at the time of ignition becomes the dominant factor; thus, the burn duration does not change significantly with the ethanol concentration. Therefore, the ethanol concentration appears to have only a small influence on 0-10% MFB at lower compression ratios and no apparent effect at higher CR. In Figure 9, it is shown that there is no common trend in the 10-90% burn duration with respect to the ethanol concentration and compression ratio. Factors such as in-cylinder mixture motion turbulence, flame quenching, dilution, and heat loss dominate this phase of the combustion process. In this case, the improvement in laminar flame speed as the ethanol concentration increases is insignificant compared to these other factors. It is also shown in Figure 9 that there is a somewhat different trend in 10-90% MFB duration with respect to the compression ratio than that of the 0-10% MFB. The 10-90% MFB duration increases as the compression ratio increases from 8:1 to 14:1, and this increase is likely the result of the surface/volume ratio becoming increasingly unfavorable to high-temperature heat loss. This is likely due to interactions between the piston and flame front, causing quenching and a reduced flame area. When Figures 8 and 9 are looked at together, it is evident that the bulk of the combustion process at MBT is phased later with an increased compression ratio. Combustion Phasing Effects. This work is the first part of a research program to complete a comprehensive study of the effects of ethanol, compression ratio, and dilution on the engine performance. As such, to minimize the need to perform complete spark sweeps during future experimentation, it was desirable to characterize the effects of the ethanol and compression ratio on MBT (peak efficiency) combustion phasing. The metric being used here for combustion phasing is CA50, which is the crank angle location where 50% of the total fuel air mixture has burned. The relationship between MBT CA50 and ethanol concentration is shown in Figure 10, where it is clear that the ethanol concentration has little influence on MBT combustion phasing. However, it is evident that the compression ratio tends to shift

MBT combustion phasing later in the cycle. This shift is likely caused by the increased surface/volume ratio of the combustion chamber at higher compression ratios. This increased surface/ volume ratio leads to higher rates of heat loss to the enginecooling system. To counteract the potential of greater heat loss, the engine favors a more retarded combustion process, thus lowering the working temperature in the cylinder, even though kinematic losses are increasing because of the reduced expansion potential. In effect, a rebalancing takes place between the competing loss functions of heat transfer and expansion potential. Despite this shift in MBT combustion phasing with CR, it is useful to note that engine performance as a function of combustion phasing is relatively constant near MBT, and because MBT combustion phasing only shifted by approximately 1.5° for a compression ratio shift of 4 points, one may conclude that a CA50 timing of approximately 12° ATDC will provide a reasonably close approximation of MBT for this engine. Conclusion In this work, MFB data was examined to study the effects of ethanol-blended fuels on IC engine combustion. This was performed in conjunction with the compression ratio as an experimental variable, while ignition timing was varied to ensure that optimal combustion phasing was attained for each set point. The following is a summary of the conclusions resulting from this study: (1) Adding ethanol to gasoline caused the KLCR to increase. KLCR for this engine is 8 for UTG 91 as compared to 16 for E84 at an engine load of 330 kPa NIMEP. (2) Increasing the ethanol fraction in the fuel mixture shows a small decrease in the 0-10% MFB duration at low compression ratios. (3) The duration of MFB during early combustion (0-10% MFB) is dominant by the effects of the compression ratio. The temperature rise of the fuel air mixture as a result of an increased compression ratio decreases the 0-10% MFB duration. With spark timing at TDC, 0-10% MFB duration decreased by 7.1 CAD from 12.55 CAD for CR of 8-5.4 CAD for a CR of 16. (4) Increasing the ethanol volume fraction from 0 to 20% decreases the bulk burn duration at a compression ratio of 8:1 and 10:1. Increases in the ethanol fraction above this volume fraction are shown to have no effect on the bulk burn duration. (5) With combustion phased at MBT, the bulk burn duration increases with an increased compression ratio. (6) At compression ratios of 8:1, MBT spark advance decreases with an increased ethanol concentration as a result of increased burn rates. At higher compression ratios, MBT spark advance remained constant because of factors other than laminar flame speed dominating burn rates. (7) The ethanol concentration has little influence on MBT combustion phasing. Increasing the compression ratio does shift MBT combustion phasing later in the cycle. Acknowledgment. The authors and Michigan Technological University acknowledge the support of General Motors (GM) for this project. In addition, support of graduate students from the National Science Foundation through the Sustainable Futures Institute IGERT project (DGE 0333401) is gratefully acknowledged. The authors also acknowledge Craig Marriot and Mathew Wiles of GM Advanced Powertrain for their discussions and input regarding this work. EF800899R