Energy & Fuels 2004, 18, 1315-1323
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An Investigation into Bioethanol Homogeneous Charge Compression Ignition (HCCI) Engine Operation with Residual Gas Trapping D. Yap, A. Megaritis,* and M. L. Wyszynski School of Engineering, Mechanical and Manufacturing Engineering, The University of Birmingham, Birmingham B15 2TT, United Kingdom Received February 8, 2004. Revised Manuscript Received May 12, 2004
Given the increasing interest in renewable fuels, bioethanol has been successfully used in conventional internal combustion engines. However, its application in homogeneous charge compression ignition (HCCI) engines requires various approaches, such as high compression ratios and/or intake charge heating, to achieve auto ignition. The approach documented here utilizes the trapping of internal residual gas (as used previously in gasoline-fueled controlled auto ignition engines) to lower the thermal requirements for the auto ignition process. In the present work, the achievable engine load range is controlled by the degree of internal trapping of exhaust gas, supplemented by moderate intake charge heating. Moderate intake heating extends the narrow auto ignition load window while varying the inlet valve timing extends the upper load range. Nitrogen oxide (NOx) emissions are characteristically low, because of the nature of homogeneous combustion.
Introduction The growing vehicle fleet around the world has intensified the search of alternative renewable fuels, which has become an important area of research, because of concerns about the potential global warming effects of major greenhouse gases from current fossil fuels. Bioethanol is considered by many as one of the most important alternatives to gasoline and diesel, because it can offer substantial reductions in the consumption of fossil fuels and the emission of greenhouse gases. Similar to ethanol, it is a pure substance (C2H5OH) that contains one O atom in its molecule. However, ethanol is produced from natural gas and crude oil, whereas bioethanol is produced from renewable raw material,1 i.e., a surplus of biomass or from crops. Bata et al.2 studied different blend rates of ethanolgasoline fuels in engines and found that the ethanol could reduce the carbon monoxide (CO) and unburned hydrocarbon (UHC) emissions, to some degree. The reduction of CO emissions is apparently caused by the wide flammability and oxygenated characteristic of ethanol. Taljaard et al.3 studied the effects of addition of oxygenates in gasoline on engine exhaust emissions and performance using a single-cylinder, four-stroke, * Author to whom correspondence should be addressed. E-mail address:
[email protected]. (1) Wang, M.; Saricks, C.; Santini, D. Effects of Fuel Ethanol Use on Fuel-Cycle Energy and Greenhouse Gas Emissions. Technical Report, Contract No. W-31-109-ENG-38, Center for Transportation Research, Energy Systems Division, Argonne National Laboratory, Argonne, IL, 1999. (2) Bata, R. M.; Elond, A. C.; Rice, R. W. Trans. ASME 1989, 111, 421-431. (3) Taljaard, H. C.; Jordaan, C. F. P.; Botha, J. J. SAE Tech. Pap. Ser. 1991, 910379.
spark ignition (SI) engine. They concluded that oxygenates significantly decreased the CO, NOx, and hydrocarbon (HC) emissions at stoichiometric air-fuel ratios. In recognition of this, there is now increasing pressure for vehicle manufacturers to produce flexible fuel vehicles that can run on E85, which is an ethanol blend (85% ethanol and 15% gasoline). Trials of bioethanolfueled buses, to exploit the low emissions of the fuel, are also currently underway in Sweden. In parallel to this interest in renewable fuels, there has also been increased interest in HCCI combustion. HCCI engines, which are also known as controlled auto ignition (CAI) engines, are being actively developed for internal combustion engines, because they have the potential to be highly efficient and produce low emissions. They can have efficiencies similar to those of diesel engines, with low emission levels of NOx and particulate matter (PM). In addition, HCCI engines have been shown to operate with a range of fuels, e.g., natural gas and gasoline.4,5 Large portions of the previous work done on fourstroke gasoline HCCI engines were conducted with high compression ratios (CRs), typically in the 15-21 region, and/or with intake air heating. Najt and Foster6 showed that HCCI combustion could be achieved in an SI fourstroke engine under lean fueling and elevated inlet charge temperatures (∼300-500 °C). The effect of increasing inlet charge temperature is to advance auto ignition timing and decrease combustion duration.6,7 (4) Christensen, M.; Johansson, B.; Einewall, P. SAE Tech. Pap. Ser. 1997, 972824. (5) Christensen, M.; Hultqvist, A.; Johansson, B. SAE Tech. Pap. Ser. 1999, 1999-01-3679. (6) Najt, P. M.; Foster, D. E. SAE Tech. Pap. Ser. 1983, 830264. (7) Thring, R. H. SAE Tech. Pap. Ser. 1989, 892068.
10.1021/ef0400215 CCC: $27.50 © 2004 American Chemical Society Published on Web 07/14/2004
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The auto ignition process has a tendency to have very rapid rates of heat release, leading to violent combustion with a very rapid rate of pressure rise; therefore, charge dilution was provided in the form of excess air (very lean air:fuel ratios) or by external exhaust gas recirculation (EGR). This dilution effectively slows the rate of combustion.6-10 Previous work with ethanol on an engine with a CR value of 11.5 and intake air heating reaching a temperature of 320 °C showed the viability of ethanol for use as an HCCI fuel.11 The engine operating range was increased, both in terms of EGR tolerance and lambda variation, in comparison to that of gasoline, coming in second only to methanol. With higher CR values of 17 and 19, intake temperatures could be reduced to 110 and 100 °C, respectively.12 Because ethanol is intended for use as a replacement fuel for gasoline, it is impractical to have such conditions for ethanol HCCI. The CR value used in gasoline-fueled SI engines is usually moderate and no intake air heating is used, making it challenging to achieve the high incylinder temperatures that are required for auto ignition. Internal trapping of residual exhaust gas has been proposed as a viable method to increase in-cylinder temperatures.13 Recent work has utilized internal residual gas recirculation for gasoline HCCI operation through different valve lifts, durations, and valve times. Lavy et al.13 and Oakley et al.14 have shown that residual gas trapping provides a charge heating effect, thus reducing the required intake temperatures. This is due to the higher temperature of residual gas in the cylinder mixing with the fresh charge, which can then be compressed to the required temperature for auto ignition. As residual rates increase, in-cylinder charge temperature also increases, resulting in an advanced start of ignition. With residual gas trapping, the rates of heat release are considerably slower, because of the significantly larger amounts of exhaust gas retained in the cylinder, as the exhaust gas acts as a diluent.15 Residual gas trapping was shown to be an effective method to achieve controlled HCCI operation for gasoline in the 4-SPACE project (where it was called controlled auto ignition (CAI)).13,16,17 One additional advantage of using residual gas trapping is that engine output/emissions can be very similar to those of stoichiometric combustion, because there is no need for excess air dilution. This allows the use of standard three-way catalytic converters that are suitable for stoichiometric combustion engine operation. (8) Gray, A. W.; Ryan, T. W. SAE Tech. Pap. Ser. 1997, 971676. (9) Ryan, T. W.; Callahan, T. SAE Tech. Pap. Ser. 1996, 961160. (10) Nakano, M.; Mandokoro, Y.; Kubo, S.; Yamazaki, S. Int. J. Eng. Res. 2000, 1 (3), 269-279. (11) Oakley, A.; Zhao, H.; Ladommatos, N.; Ma, T. SAE Tech. Pap. Ser. 2001, 2001-01-3606. (12) Christensen, M.; Johansson, B.; Amne´us, P.; Mauss, F. SAE Tech. Pap. Ser. 1998, 980787. (13) Lavy, J.; Dabadie, J. C.; Angelberger, C.; Wiland, J.; Juretzka, A.; Schaflein, J.; Ma, T.; Lendresse, Y.; Satre, A.; Schulz, C.; Kramer, H.; Zhao, H.; Damiano, L. SAE Tech. Pap. Ser. 2000, 2000-01-1837. (14) Oakley, A.; Zhao, H.; Ladommatos, N. SAE Tech. Pap. Ser. 2001, 2001-01-1030. (15) Zhao, H.; Peng, Z.; Williams, J.; Ladommatos, N. SAE Tech. Pap. Ser. 2001, 2001-01-3607. (16) Li, J.; Zhao, H.; Ladommatos, N. SAE Tech. Pap. Ser. 2001, 2001-01-3608. (17) Zhao, H.; Li, J.; Ma, T.; Ladommatos, N. SAE Tech. Pap. Ser. 2001, 2001-01-0420.
Yap et al. Table 1. Engine Specification Summary engine type bore stroke compression ratio fueling type
Medusa single-cylinder 4-V engine 80 mm 88.9 mm 12.5 liquid port-injected
Bioethanol has a higher auto ignition temperature than gasoline, which is reflected in its high research octane number (RON), and it has been used extensively in gasoline blending to improve its knock resistance. The high auto ignition temperature requirement for bioethanol makes it a suitable candidate for the application of residual gas trapping, to utilize the charge heating effect to lower the intake charge heating requirements while keeping CRs at moderate levels. In addition, due to the nature of HCCI combustion, low NOx emissions will be reduced in comparison to ethanol SI engines. This paper presents the application of this strategy to ethanol HCCI and the useable operating range of the engine by varying the inlet and exhaust valve timing. This allows for varying amounts of trapped residual gases, which will also vary the engine load. Experimental Section Engine Setup. A modified Medusa single-cylinder engine was used to examine the effect of valve timing on engine load and residual gas trapping. The engine was coupled to a dc dynamometer, which kept the engine at a constant set speed. The CR value was increased from the standard specification of 10.5 to a value of 12.5, using a racing-style piston. This adaptation allowed the combustion chamber shape to be more similar to the typical advanced engine designs used for future HCCI operation. Bioethanol fueling was performed via a standard injector located close to the inlet port of the engine. A summary of the engine specifications is given in Table 1. A Kistler model 6125A pressure transducer was fitted flush with the wall of the combustion chamber, which was connected via a Kistler model 5011 charge amplifier to a National Instruments data acquisition board that was installed in an IBM-compatible personal computer (PC). The crankshaft position was measured using a digital shaft encoder. Software was developed in-house, in the LabVIEW programming environment, to record the in-cylinder pressure versus crank angle for 100 consecutive engine cycles, and to analyze the resulting data. Output from the analysis of consecutive engine cycles included peak engine cylinder pressure, values of the indicated mean effective pressure (IMEP), the percentage coefficient of variation (% COV) of IMEP, average values and percentage COV of peak cylinder pressures, average crank angle for ignition delay, burn duration, 5% and 95% burn point, etc. The engine was operated at stoichiometric air:fuel ratio. The valve timings, engine IMEP, and pressure traces were recorded for calculation of the trapped residual fractions. Carbon dioxide (CO2), carbon monoxide (CO), unburned hydrocarbons (UHCs), oxygen, and nitrogen oxide (NOx) emissions were also recorded. All the tests presented here were conducted at an engine speed of 1500 rpm. A 3 kW electric air heater was installed in the intake duct to preheat the air to extend the useable load range of bioethanol. The heater was located upstream of both the fuel entry port and the exhaust gas recirculation (EGR) loop. The intake temperature was measured in the intake port after the fuel injector, ∼70 mm from the intake valve seats, and it was used to control the heating power. Anhydrous bioethanol (supplied by Shell Global Solutions, U.K.) was used for all the tests that have been presented. Valve Timing Setup. Special valve strategies were used to trap residual gases and vary the engine load. The engine
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Figure 1. Calculation method for trapped residuals. throttle was kept wide open throughout the tests. The inlet and exhaust valves were set manually before operation with a vernier-adjusted camshaft pulley. The exhaust valve timing determined the engine load, whereas the inlet valve was swept to examine the effect of the inlet valve timing for set exhaust valve timings. The range was limited by the sweep of the verniers. Proprietary fixed duration camshafts were used. This procedure allowed control of the amount of residual gas, thus controlling the amount of the freshly inducted charge and, hence, varying the achievable load. The lower achievable engine load limits were decided by the occurrence of misfiring of the engine when in SI combustion during startup. To achieve HCCI combustion without variable valve timing, the required valve event had to be set before starting the engine in SI. For HCCI operation, at advanced exhaust valve timings, misfiring of the engine during startup was common, hence limiting the lower load achievable. Because electronically variable valve timing was not available, the data for the lower boundary of HCCI operation admittedly is expected to be limited by the SI misfiring and, hence, might not be the true lower limit of HCCI combustion. The upper load limit was decided by the knock limit or the misfiring limit of the engine. The valve timing parameters used in this paper were as follows: (1) The maximum opening point (MOP) of the inlet and exhaust valves are given. (2) The inlet valve MOP is given in crank angle degrees (CAD) after exhaust stroke top dead center (TDC). (3) The exhaust valve MOP is given in CAD before exhaust stroke TDC. The inlet valve MOP was swept between 118 CAD to 162 CAD after TDC (or 478 CAD to 522 CAD, referring to the 720 CAD engine cycle with 0 CAD set at the firing TDC), whereas the exhaust valve MOP was swept between 142 CAD to 162 CAD before TDC (or 218 CAD to 198 CAD). Engine data acquisition and analysis was performed at spacings of every 4 CAD for the inlet and exhaust valve sweep. Calculation Procedures for Trapped Residual Fraction. Trapped residual gas volumes were calculated as a fraction of the total volume of the cylinder using the following method, referring to Figure 1. The residual gas fraction was calculated from the ratio of the cylinder pressures at two points: one point after combustion and the other at the same volume (same offset from TDC) during the compression of the residual gas, raised to the power of an empirical factor (1/K), derived from the isentropic compression and expansion theory, with special considerations for heat transfer during this period. This heat-transfer factor
was determined and validated by a combination of simple thermodynamic calculations for gas-exchange and in-cylinder heat-transfer modeling, using Ricardo WAVE engine software. Through statistical error propagation, accurate pressure readings were determined to affect calculations more significantly than heat-transfer errors; however, because of the low combustion temperatures and short burn duration that are typical of HCCI combustion, thermal shock was believed to be of minimal effect.
Results and Discussion Engine Load Range. Figure 2 is a map of the engine loads achievable for the range of inlet and exhaust valve timings used. The lines represent contours of the load map created from the engine data acquisitions at spacings of every 4 CAD for the inlet and exhaust valve sweep, each of the contours having markers to show the engine load achievable for that particular area. Without any intake air heating, the load range of the engine with bioethanol was 2.4-3.4 bar IMEP. This is admittedly much smaller than that of an SI engine. However, naturally aspirated engines fueled with other fuels (such as gasoline) have similar load ranges when running in HCCI operation.17 The highest load can be achieved at an inlet valve MOP of 138 CAD, whereas the exhaust MOP is set at 150 CAD. Further retarding of the exhaust valve failed to increase the engine load, because the combustion became unstable. This was due to the reduction of trapped residuals, which would reduce the heating effect within the cylinder. As a result, there would be insufficient thermal energy to bring the fuel up to auto ignition temperature. It is also apparent that the higher engine loads are obtained when the inlet valve MOP is smaller in magnitude, as compared to the exhaust valve MOP. It is more prominent when the exhaust valve timing is 150-154 CAD, where the inlet valve timing needs to be set at 130-146 CAD for stable HCCI combustion. The cylinder volume at which the inlet valve MOP occurs after the gas-exchange TDC is smaller than the cylinder volume at which the exhaust valve MOP occurs before the gas-exchange TDC. For ease of reference, the inlet valve MOP, for this case, shall be referenced as being advanced, compared to the exhaust valve MOP about the gas-exchange TDC.
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Figure 2. Engine load range (IMEP, in bar) without intake air heating.
Figure 3. Plot showing the 5% burn point (CAD) without intake air heating.
When the inlet valve is retarded, in comparison to the exhaust valve MOP about the gas-exchange TDC, the engine load will start to decrease until reaching a point at which combustion will not be sustained. For significantly retarded inlet valve MOP timing, the engine would misfire. The 5% burn point of the combustion shows the position of the 5% burn point of bioethanol, in terms of CAD, where a negative number denotes that it occurs in advance of TDC and a positive number means that it has occurred after TDC. In Figure 3, the dotted line represents the path where the start of combustion is most advanced for the range of exhaust valve MOP timings. Following the dotted line, it can be observed
that, for any given exhaust valve MOP, the 5% burn point is most advanced when the inlet valve is ∼10-15 CAD advanced, as compared to the exhaust valve MOP about the gas-exchange TDC. The 5% burn point is a good indication of the start of combustion, and, from the graph, it can be deduced that the start of combustion is most advanced when the inlet valve is slightly advanced, in comparison to the exhaust valve. Because the combustion process in homogeneous auto ignition has no external governing event, such as that of a spark plug in an SI engine, the start of combustion is a function of pressure, temperature, and fuel properties. There are empirical correlations that can be used
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Figure 4. Change in compression ratio with varying inlet valve timing.
to estimate the start of combustion from ignition delay, and one such correlation18 is shown below:
τ ) Ap-neB/T
(1)
where τ is the ignition delay; p and T are the cylinder pressure and temperature, respectively; and A, B, and n are positive coefficients that are related to the fuel and are determined experimentally. As eq 1 shows, a shorter ignition delay and, hence, a more advanced start of combustion can be achieved if the temperature is higher, if there are favorable fuel properties and, to a smaller extent, pressure. Because the intake air was not heated, the intake temperatures in the port of the engine did not change significantly at the various points when the start of combustion was most advanced. There was no artificial elevation of the temperature, and the fuel properties were unchanged because the same stock was used for all the tests. However, the camshafts used have fixed duration; therefore, varying the maximum opening point of the inlet valve would also vary the inlet valve closure. Figure 4 shows the calculated CR of the engine from the crank volume at inlet valve closure (IVC). As shown, by varying the IVC, the CR of the engine changes. There is a minimal change from 118 CAD to ∼140 CAD; however, as the inlet valve MOP approaches 162 CAD, there is a much greater change in the CR value. As the inlet valve MOP approaches 162 CAD, the IVC occurs when the piston on the upstroke is accelerating. Hence, a small change in angle results in a larger reduction in chamber volume at IVC. A higher compression ratio would result in a higher end-gas temperature at TDC before ignition, thus reducing the ignition delay. A lower CR value from a retarded IVC would then result in a lengthened ignition delay. However, if the inlet valve is advanced too much, in comparison to the exhaust valve MOP about the gasexchange TDC, there is still a high cylinder pressure when the inlet valve opens. This is because the inlet valve opening (IVO) is also varied as the inlet valve MOP varies, thus occurring much earlier. This would (18) Ohyama, Y. SAE Tech. Pap. Ser. 2000, 2000-01-0198.
then result in the hot residual gases escaping from the cylinder into the inlet manifold due to pressure difference and cooling. Zhao et al.17 called this “early backflow”. Examination of the intake air temperatures, as given in Figure 5, shows that, at significantly advanced inlet valve timings, the intake air temperature starts to increase. The graph shows that as the inlet valve timings approach 118 CAD, the intake air temperature increases, from 23 °C up to 53 °C. These temperature measurements were taken in the intake port of the engine close to the inlet valves. The inlet is not heated; therefore, this increase in temperature would be the result of the backflow of trapped residuals into the inlet. Figure 3 shows that, from 130 CAD down to 118 CAD, as the intake temperature increases, the 5% burn point and, hence, the start of combustion is gradually retarded. This would then indicate a reduction of the in-cylinder thermal energy that is due to the loss of it from the backflow into the intake manifold. Hence, in this engine setup, it seems that, to maximize the heating effect within the cylinder, the inlet valve timings must be advanced in comparison to the exhaust valve timing, because of a balance between a reduction in CR and backflow from trapped residual gases into the intake manifold. Without intake air heating, it seems that the maximum engine load could only be reached for a small set of inlet valve and exhaust valve timings. Some intake air heating was then applied to determine if the load range window could be enlarged. An intake air temperature of 80 °C was used. This temperature was selected because it was the temperature of the coolant in the engine and it might be possible in a production-style engine without the use of auxiliary heating equipment. The intake charge was heated for an exhaust valve timing of 150 CAD and for subsequent exhaust valve timings, which were more retarded. As indicated by Figure 6, which shows the engine load range with intake air temperature of 80 °C, it is now possible to extend the operating window of the engine with retarded exhaust valve timings. In addition, the maximum possible engine load is now 4.18 bar IMEP. This was achieved at an exhaust valve timing of 142 CAD and an inlet valve timing of 118 CAD. Combustion Indicators. Figure 7a shows that, without intake air heating, a large proportion of the engine load range has a maximum rate of pressure rise of
