ARTICLE pubs.acs.org/EF
Lean Combustion and Emission Characteristics of Bioethanol and Its Blends in a Spark Ignition (SI) Engine Seung Hyun Yoon† and Chang Sik Lee*,‡ †
Department of Mechanical Engineering-Engineering Mechanics, Michigan Technology University, 917 R.L. Smith Building, 1400 Townsend Drive, Houghton, Michigan 49931, United States ‡ Department of Mechanical Engineering, Hanyang Univeristy, 17 Haengdang-dong, Sungdong-gu, Seoul 133-791, Korea ABSTRACT: Lean combustion and exhaust emission characteristics in a spark ignition engine (SI engine) with variation of the ethanol gasoline blending ratio and the excess air ratio were investigated in this research. To investigate the influence of the excess air ratio and ethanol blends, the lean combustion characteristics such as brake torque, cylinder pressure, and the rate of heat release (ROHR) were analyzed under the various excess-air ratios. In addition, the reduction effects of exhaust emissions such as carbon monoxide (CO), unburned hydrocarbon (HC), oxides of nitrogen (NOx), and carbon dioxide (CO2) were compared with those of gasoline fuel. The results showed that the peak combustion pressures and the ROHR of all test fuels linearly decreased as the excess air ratio (λ > 1.0) increased. As compared with gasoline fuel (G100) at each given excess air ratio, there were slight improvements in combustion pressure for ethanol blended fuels (E20 E100). The power output and brake mean effective pressure (BMEP) slightly increased at each air fuel ratio condition compared to G100 as the increase of ethanol fraction. The difference in the power and BMEP between E100 and G100 were maximized with the increases in the air fuel ratio up to λ = 1.5. Ethanol blends have higher BSFCs compared to G100 and also achieved fairly stable combustion features at all excess air ratios compared to gasoline. NOx emissions tended to decrease proportionally with increases in the excess air ratio for all test conditions, and all of the ethanol blends emitted slightly less NOx compared to G100.
1. INTRODUCTION In recent years, given the dramatically increasing number of cars, public concern has steadily increased regarding a possible shortage of fossil fuel resources, energy safety policies, and environmental pollution regulations. In particular, the environmental problems and growing energy crises associated with global warming and energy preservation have enhanced the focus on alternative fuel research and the fuel efficiency of the internal combustion engine.1,2 With regard to alternative fuels, ethanol and ethanol gasoline blended fuels are considered to be most promising potential substitutes for petroleum gasoline in conventional spark ignition (SI) engines. Bioethanol is a biomass-based renewable, biodegradable, and environmentally friendly alternative fuel because it can be produced from cellulosic biomass, agricultural feedstock, and scrapped resources such as corn, wheat, sugar cane, plants, and wood waste through distillation and fermentation processes. It is also the most feasible alternative fuel since it can be easily implemented into current refueling infrastructures. Given these numerous benefits, bioethanol and ethanol gasoline blended fuels are becoming more widely used in automotive vehicles and garnering increased investigation.3 5 The combustion of ethanol in conventional SI engines demonstrates that improved brake thermal efficiency can be achieved by increasing the compression ratio without producing knock-phenomena because of the high octane number of ethanol fuel. Furthermore, harmful exhaust emissions of greenhouse gases and unburned HC and CO can also be reduced because of the complete and stable combustion phases of the hydroxyl group and oxygen compounds (approximately 35%) in ethanol. r 2011 American Chemical Society
In addition, the high latent heat of vaporization permits a higher volumetric efficiency compared to that of gasoline fuel due to the evaporative cooling effect of intake charge.6,7 Due to the advantages of bioethanol and its blends with gasoline fuel, much research regarding ethanol fuels has focused on automotive SI engines. Al-Hasan8 reported on the effects of a gasoline ethanol blend on combustion and emission characteristics. The results revealed that fuel blends increased the brake power, torque, volumetric and brake thermal efficiencies, and fuel consumption rate with equivalent air fuel ratios. The concentrations of CO and HC emissions from ethanol combustion were significantly decreased compared to those from gasoline combustion. Hsieh et al.9 experimentally investigated the performance and exhaust emissions of an SI engine fueled with ethanol gasoline blends with ethanol fractions ranging from 0% to 30%. The experimental results showed that the engine torque output and brake specific fuel consumption (BSFC) of ethanol gasoline blended fuels slightly increased with increases in blending fraction and that the CO and HC emissions decreased due to the lean effect of ethanol. Despite the advantages of ethanol, it has faced some challenges in becoming widely implemented for automotive use and in industrial power production applications. The latent heat of vaporization for ethanol is higher than gasoline and its low vapor pressure and high boiling point generally cause poor performance in cold conditions.10,11 More specifically, the lower Received: May 5, 2011 Revised: July 3, 2011 Published: July 05, 2011 3484
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Table 1. Properties of Test Fuels property
gasoline
ethanol C2H5OH
chemical formula
CnH1.87n
molecular weight (kg/kmol)
114.15
46.07
oxygen (%wt)
0
35
octane number (RON)
86 94
105 108
boiling point (°C)
30 180
78.5
latent heat of vaporization (kJ/kg)
289
854
autoignition temperature (°C)
257
423
A/F ratio (by volume) lower heating value (MJ/kg)
14.7 43.8
9.0 26.7
heating value (LHV) and stoichiometric air fuel ratio of ethanol are relatively lower compared to those of gasoline. To achieve the same power or torque as gasoline, a larger amount of ethanol fuel must be supplied to the combustion chamber, resulting in a higher fuel consumption rate.12 Accordingly, the use of lean mixtures of ethanol fuel in SI engines will enable a feasible way to improve fuel consumption and thermal efficiency while lowering exhaust emissions because of the better combustion efficiency of the in-cylinder mixture and the reduced combustion temperature at optimal lean conditions.13 Therefore, research is gradually increasing with regard to lean mixture combustion and lean operational mixture range for ethanol fuel and ethanol gasoline blended fuel. The objective of the present work was to investigate the lean operational limits and the effects of variation of the excess air ratio on the combustion performance and reduction of exhaust emission characteristics in an SI engine fueled with ethanol, ethanol gasoline blended fuels, and gasoline. As such, the combustion pressure and the rate of heat release (ROHR) were evaluated and the cycle-to-cycle variations of combustion were analyzed under changing operating conditions, including varying excess air ratios and mixing ratios of ethanol, in order to analyze the combustion and emission characteristics of the test fuels in an SI engine. Additionally, the test fuels were investigated under various engine conditions to compare engine performance, combustion stability, fuel consumption, thermal efficiency, brake torque, brake power, brake mean effective pressure (BMEP), and the concentrations of NOx, HC, CO, and CO2 emissions. Physical properties such as density, viscosity, surface tension, and distillation rate were also investigated according to variations in ethanol fraction and fuel temperature.
2. EXPERIMENTAL APPARATUS AND PROCEDURE
Figure 1. Physical properties of test fuels at specified fuel temperatures.
2.1. Test Fuels. In this study, the reference fuel was an unleaded gasoline without any additives (denoted G100 hereafter), which satisfies the ASTM standard specification. Bioethanol (denoted E100 hereafter) with an anhydrous ethanol purity ratio of 99.9%, which meets the ASTM D5798-09 standard,14 was also used. The fractions of ethanol for the ethanol gasoline blended fuels were 20%, 40%, 60%, and 80% (denoted E20, E40, E60, E80 hereafter, respectively), which means the volumetric percentage of ethanol fuel. Table 1 and Figure 1 show the fuel properties of the gasoline and bioethanol used in this experiment. The properties of E100 are different compared to those of G100. In particular, E100 is an oxygenated fuel and therefore contains an oxygen molecule (approximately 34.7% by vol.) in the fuel. Considering the chemical formula (E100 C2H5OH,
G100 CnH1.87n, n = 4 12), E100 has a lower heating value (about 26.7 MJ/kg) and stoichiometric air fuel ratio (9.0) compared to gasoline (approximately 43.8 MJ/kg and 14.7, respectively). With regard to research octane number (RON), E100 has a higher octane number (105 108) than gasoline (86 94). Furthermore, E100 has a higher latent heat of vaporization (approximately 854 kJ/kg) compared to G100 (approximately 289 kJ/kg). Figure 1 shows the measured physical properties of density, kinematic viscosity, and surface tension according to bioethanol fraction and fuel temperature by using a gravimeter, a viscometer, and a surface tension measurement system. As the fuel temperature is increased from 0 to 70 °C, the density, viscosity, and surface tension linearly 3485
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Table 2. Specifications of Test Engine engine type bore
four-cylinder SI engine 77.0 mm
stroke
85.44 mm
displacement volume
1591.0 cm3
compression ratio ignition system
10.5 DLI (distributor less ignition)
ignition sequence
1 3 4 2
maximum power
89.0 kW @ 6200 rpm
maximum torque
15.6 kgf 3 m @ 4200 rpm
valve timing
Figure 2. Distillation rates for test fuels according to fuel temperature.
Figure 3. Schematic diagram of experimental apparatus.
decrease. Furthermore, as the ethanol fraction is increased, the density, viscosity, and surface tension linearly increase due to the addition of ethanol. The experimental results for distillation of the test fuels in compliance with ASTM D86-08a15 are shown in Figure 2. The initiation of distillation of G100 appeared at a lower temperature compared to the blends and E100 due to its lower boiling point (30 180 °C); furthermore, distillation rapidly increased with increases in fuel temperature. However, for E100, distillation began at a boiling point of 78.5 °C and continued over time. As the ethanol fraction decreased, the initial distillation point rose and increased linearly with the increase in fuel temperature, which indicates that the distillation rates of the blends were correlated with the evaporation of gasoline fuel. These different fuel properties may affect fuel atomization and evaporation characteristics in the combustion chamber during the combustion process. 2.2. Experimental Apparatus. The experimental system was composed of a test engine, engine control system, dynamometer controller, combustion analyzer, and emissions measurement systems as shown in the schematic diagram in Figure 3. The test engine used for this study was an electronic fuel injection (EFI), 4-cylinder, 16-valve SI engine with a sweep volume of 1591 cm3 (L). It had a bore of 77.0 mm, a stroke of 85.44 mm, and a compression ratio of 10.5. The maximum generated engine power and maximum engine torque were approximately 89 kW
IVO
10° ATDC
IVC
63° ABDC
EVO EVC
40° BBDC 38° ATDC
at 6200 rpm and 152.88 N m at 4200 rpm, respectively. The detailed specifications and dimensions of the test engine are summarized in Table 2. The engine output and loads were measured and controlled using an eddy current (EC) dynamometer system (EC-80, MEIDEN) with a maximum braking power of 105 kW. To control the operating conditions, components such as the engine actuators, fuel injector, ignition coil, and idle speed control valve were coupled to an engine management system (M8, MOTEC). This allowed the injector pulse width (injection mass), injection timing, and spark ignition timing to be precisely adjusted to the desired test conditions. To appropriately adjust the EMS map datum, the engine management system also monitored the crank angle, the absolute pressure of the intake manifold (air flow rate), and the throttle valve opening. The flow rate of the intake-air and the temperature of the intakeair, exhaust gas, and coolant were monitored by the data acquisition system. The injection duration (injection mass) was adjusted using a closed-loop lambda control system in order to evaluate the effects of the various test fuels and variation of the air fuel ratio. The air fuel ratio was controlled by the EMS coupled to a lambda meter (PLM, MOTEC) with a feedback signal from a wide-band oxygen sensor (Bosch, LSU 4). The incylinder (combustion) pressure was measured using a piezoelectric pressure sensor (6052B1, Kistler) and a measuring spark plug (6117BFD17, Kistler). These were coupled to a charge amplifier (5011B, Kistler) and data were recorded with the crank angle degree. The combustion pressure data were measured and recorded using a data acquisition board (PCI-MIO-16E-1, National Instrument) with a maximum sampling rate of 1 megasample/sec and a developed combustion acquisition and analysis program to ensure accurate ignition timing and phasing of heat release. The obtained in-cylinder pressure data versus the crank angle were averaged over 300 engine cycles to reduce the effect of cycle-to-cycle variations; the rate of heat release (ROHR) was then calculated for the analysis of the combustion characteristics for each test fuel and set of conditions. For the case of emissions measurements, the exhaust emissions from the engine were measured with a NOx analyzer with measuring accuracy of less than (1.0% (BCL-511, Yanco), and a HC CO emissions analyzer with the repeatability of (12 ppm HC, ( 0.06% vol. CO (Mexa-554JK, Horiba). Measurements of the exhaust emissions were conducted after precalibration of the emission analyzers and once the test engine had sufficiently stabilized to steady-state engine (temperature) 3486
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conditions. The consumption of fuel was measured using a fuel flow meter system (Froude, FG100), which has a 0 200 L/min flow range, accuracy of (1%, and repeatability of (0.5% with a given pressure. Table 3. Experimental Conditions test fuels excess air ratio (λ)
G100, E100, E20, E40, E60, E80 1.0 - 1.5
engine speed (rev/min)
3000
engine load (%, TP)
20
ignition timing (deg. BTDC)
MBT timing
injection timing (deg. BTDC)
40
intake air temperature (°C) coolant temperature (°C)
30 70
2.3. Experimental Conditions. The experimental conditions for this study are listed in Table 3 and as follows. The test fuels were E100, E80, E60, E40, E20, and G100, the names of which indicate the content of ethanol in different volume ratios (e.g., E20 is composed of 20% ethanol and 80% gasoline by volume). To investigate the effect of the excess air ratio (λ) on the combustion and exhaust emission characteristics, the engine test parameters for the lean combustion limit were set to define the lean combustion limits with the test fuels. Combustion around the lean limit can cause unstable combustion phenomena in the engine including misfiring, combustion variation, or a sharp increase in unburned emissions. The experiments were performed at various air fuel ratios (λ) ranging from 1.0 up to 1.5 in 0.1 intervals. The relative air fuel ratio (λ) was defined as (AFR)act, the actual air fuel ratio, over (AFR)st, the stoichiometric air fuel ratio for each test fuel. All engine tests were
Figure 4. Combustion characteristics of test fuels with varying excess air ratios. 3487
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Energy & Fuels conducted at fixed engine speed conditions of 3000 rpm and were maintained by the EC dynamometer using a constant engine speed mode. The test engine load was fixed at 20% (part load) and the injection strategy, including injection duration (injection mass), was based on fuel type and modified to maintain a set lambda (λ) value for each fuel blend. Using this method the engine load could be adjusted by opening the throttle valve to 20%; the rated brake torque corresponding to each operating condition was measured with the EC dynamometer. To determine the spark timing required for maximum brake torque (MBT) for each test fuel, the spark ignition timing was manually controlled using the ECU control and modifying the top dead center (TDC), before top dead center (BTDC), and crank angle (CA; started at 40° and altered at 5° intervals). To ensure repeatability and reduce variations in the measurements of the test fuels and conditions, the coolant temperature and lubricant oil temperature were kept at 70 ( 2 °C for all tests. The intake air temperature was also fixed at 30 ( 1 °C. The injection timing was kept at BTDC 40 deg.
3. RESULTS AND DISSCUSSION The effects of variations in the excess air fuel ratio on the incylinder pressure (MPa) and rate of heat release (ROHR, J/deg) according to the blending ratio of the test fuels, with constants of engine speed at 3000 rpm, MBT timing, and throttle opening at 20%, are shown in Figure 4a f. In the pressure-crank angle diagram for each fuel, the maximum combustion pressures appeared with a stoichiometric air fuel ratio (λ = 1.0). As the excess air ratio (λ > 1.0) increased, the peak values for the maximum combustion pressure and ROHR of the test fuels linearly decreased. Thus, the minimum combustion pressures and ROHR were generated at λ = 1.5. The total amount of fuel injection became insufficient to complete combustion, as compared with induced intake air mass at lean conditions. As a result, the output performance decreased as the excess air ratio increased. However, there were slight improvements in the combustion pressures for the ethanol blended fuels (E20 E100) when compared with G100 at each given excess air ratio because the heat evaporation rate of ethanol (854 kJ/kg) is 2.96 times higher than that of gasoline fuel (289 kJ/kg) and because the MBT timing of ethanol advanced with a high octane number, as shown in Table 1. The effects of a higher latent heat on the increased intake air density and induced air charge results in an increased fuel mass injected into the combustion chamber during the intake process. In addition, ethanol blended fuels provide additional oxygen for combustion performance due to the oxygen content of the fuel. More specifically, Figure 5 demonstrates how the trend in peak combustion pressure changed as the ethanol fraction decreased. As the bioethanol blending ratio is increased, the combustion peak pressure is increased according to the decrease of excess air ratio. In the case of E100 fuel, the peak pressure is higher than the results of G100 fuel as illustrated in Figure 5. This trend can be explained by the fact that the increasing latent heat with the increase in ethanol fraction tends to decrease the temperature of the intake air, which reduced the compression work and resulted in the active combustion. Furthermore, it can be seen that each ethanol blend presents a narrow variation in peak pressure with the excess air ratio around λ = 1.0 1.1, while G100 combustions demonstrate wide fluctuations and reductions in combustion pressure. Compared
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Figure 5. Peak combustion pressure of test fuels with varying excess air ratios.
Figure 6. Brake power and BMEP of test fuels with varying excess air ratios.
to G100, with ethanol blends there are fewer effects of unstable combustion phenomena, such as engine knocking. For lean operation ranges (λ > 1.2), the combustion of ethanol blends showed the weak influences of undesirable and incomplete combustion because of the increased injection mass needed to compensate for the decreased low heating value (LHV). For both the ethanol blended fuels and gasoline, the effects of the air fuel ratio on the brake power (KW) and brake mean effective pressure (BMEP, MPa), with a throttle opening of 20% and an engine speed of 3000 rpm, are shown in Figure 6a and b, 3488
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Figure 7. Exhaust gas temperature of test fuels with varying excess air ratios.
respectively. The maximum brake power and maximum brake mean effective pressure of the test fuels were generated at the stoichiometric air fuel ratio (λ = 1.0), the combustion pressure characteristics were likewise optimized because the suitable air fuel charge allowed flow into the combustion chamber for complete combustion. The brake power and BMEP linearly decreased for all test fuels when the air fuel mixture shifted toward a lean condition (λ > 1.0), as expected. This is due to a decrease in the amount of fuel supplied for combustion as the intake air mass increased with the lean condition. With regard to the influence of the ethanol blended fuels on engine performance, as the ethanol fraction increased the power and BMEP increased by approximately 1.2% compared to G100 for each air fuel ratio. Additionally, the differences in power and BMEP between E100 and G100 were maximized at the highest air fuel ratio. With λ = 1.5, the maximum power and BMEP for E100 were 14.1 kW and 0.415 MPa, respectively. These values are approximately 14.6% and 15% higher than for G100 (12.3 kW, 0.361 MPa), respectively. With ethanol fuel, the MBT timing is increased by the high octane number, which results in improved combustion performance. At the same time, the high heat of evaporation of ethanol increases the density of the intake air and the volumetric efficiency of the intake process. The resulting increase in the amount of air fuel mixture flowing into the combustion chamber improves combustion performance. The high oxygen content of ethanol fuel also leads to improved combustion performance.17,18 Figure 7 shows the exhaust gas temperature, measured at the end of the exhaust port, plotted against the air excess ratio and blending fraction of ethanol fuel at an engine speed of 3000 rpm and throttle opening of 20%. For all test fuels, the maximum exhaust gas temperature occurred at the stoichiometric air fuel ratio (λ = 1.0) due to the sufficient amount of air. The gas temperature linearly decreased as the excess air ratio increased. Additionally, for each given excess air ratio, the gas temperature of the ethanol blends was lower than that of gasoline. In general, gas temperatures are reduced as the ethanol fraction is increased. This result can primarily be explained by the charge dilution and cooling effects caused by the high latent heat of evaporation and heat capacity of ethanol fuel. Furthermore, the low adiabatic flame temperature of ethanol contributes to the lower gas temperature. Figures 8 and 9 respectively indicate the effect of the excess air ratio and ethanol blends on the brake specific fuel consumption
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Figure 8. BSFC of test fuels with varying excess air ratios.
Figure 9. BSEC and brake thermal efficiency of test fuels with varying excess air ratios.
rate (BSFC), brake specific energy consumption (BSEC), and brake thermal efficiency under the conditions of constant MBT, 3000 rpm, and throttle opening of 20%. In Figure 8, the fuel consumption rate (BSFC) for all test fuels increased as the excess air ratio increased. Under a given condition, constant engine speed (3000 rpm) and throttle opening (20%), the amount of intake air is constant for each test fuel. Accordingly, the mass of fuel injection is reduced to correspond with the increase of excess air ratio. The fact that the BSFC increased as the excess air ratio increased suggests that the lower power output can be attributed to the reduced injection mass and incomplete combustion in the lean combustion region. Additionally, ethanol blends have higher BSFCs compared to G100; these higher BSFCs are related to the lower heating value and higher air intake flow rate of ethanol blends. The low heating value (LHV) of ethanol fuel is approximately 26.7 MJ/kg, which is approximately 39.1% lower than that of gasoline fuel (43.8 MJ/kg). Furthermore, ethanol fuel increases volumetric efficiency because of its cooling effects and the larger amount of fuel required in the cylinder. The stoichiometric air fuel ratio for ethanol is approximately 9 to 1, while for gasoline it is approximately 14.7 to 1. Therefore, given equal induced-air flow rates, more ethanol is required to satisfy the stoichiometric ratio compared to gasoline. On the contrary in Figure 9, in the case of BSEC considering the LHV value of test fuels, it can be seen that G100 indicates highest value, while the lowest value corresponds to E40. With the BSEC results, the conversion efficiencies of consumed fuel 3489
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Figure 10. COV of IMEP of test fuels with varying excess air ratios.
into energy of ethanol blends are much better than those of gasoline fuel and also it can achieve improved results of brake thermal efficiency. However, the thermal efficiencies of test fuels are decreased for lean combustion conditions due to the deteriorated combustion stability. The above results indicate that the conversion efficiencies for the ethanol blends was higher than it was for gasoline. Furthermore, the brake thermal efficiencies of the ethanol blends were also higher than those of gasoline; however, the fuel consumption rates of the ethanol blends were also increased compared to gasoline. To investigate the effect of lean combustion and the addition of bioethanol fuel on combustion stability, the parameters of combustion variation were investigated by analyzing the coefficient of variation (COV) of the indicated mean effective pressure (IMEP). The COVIMEP values during 300 consecutive engine cycles for all test fuels at a fixed engine speed of 3000 rpm and throttle opening of 20% condition were measured and are shown in Figure 10. The COVIMEP of all test fuels demonstrate rapidly increasing trends around an excess air ratio of 1.5. At λ = 1.5, which is an excessively lean condition, the COVIMEP of G100 approaches 13.12% while that of E100 approaches 7.94%. This indicates that the excessively lean air fuel ratio causes a deterioration of combustion stability and results in unstable combustion with cycle-to-cycle variation. However, the ethanol fuels achieved fairly stable combustion features at all excess air ratios compared to gasoline combustion. This can be attributed to the suppression of unstable combustion and the resulting low COVIMEP levels. Moreover, when the lean burn limit (LBL) is supposed to a general criteria of COVIMEP = 5% to consider the stability of combustion, E80 and E100 combustion can be extended with a LBL up to λ = 1.4. These results suggest that the operating range limit can clearly be extended by using ethanol fuel for improved combustion stability due to its antiknock quality. This is primarily caused by the high octane number and the high latent heat of evaporation for ethanol fuel. Additionally, the effect of uncompleted combustion phenomena becomes weak with the increased injection quantity that compensates for the low LHV as the lean operation range is approached. Figure 11a, b, and c show the influence of different ethanol gasoline blended fuels at different air fuel ratios on exhaust emissions such as NOx, HC, and CO. In the case of Figure 11a, the peak concentrations of NOx emissions for all test conditions exhausted at an excess air ratio of λ = 1.1, where the combustion temperature is high due to the mixture abundant in oxygen and
Figure 11. Exhaust emission characteristics of test fuels with varying excess air ratios.
fuel, and then the concentration of NOx emissions tended to decrease proportionally with the increase of excess air ratio for all test conditions. Furthermore, all ethanol blends combustion emitted slightly less NOx emissions than that of gasoline fuel, which is approximately 15% below the level measured with gasoline fuel. The lowest level of NOx emission was achieved using E100 with lean combustion conditions. These trends of ethanol and its blends are primarily due to the low combustion temperature with the high latent heat, and the oxygen content of ethanol fuel. In the combustion process, NOx emissions are generally produced from the high temperature reaction of oxygen and nitrogen. However, in the case of ethanol fuel, its high latent heat leads to a decreased temperature of the charge mixture due 3490
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Energy & Fuels to the cooling effects during the intake process in the combustion chamber. Additionally, the oxygen content of ethanol fuel provides more oxygen in the combustion chamber and results in an additional lean effect at a constant excess air ratio. Therefore, lean combustion with ethanol blends leads to a lowering of the charge mixture and combustion temperature, thus suppressing the activation of thermal NOx formation. With regard to HC emissions, there was no obvious increase in HC emissions with conditions ranging from the stoichiometric ratio (1.0) to λ = 1.3, at which point sufficient air exists to allow unburned HC emissions to participate in the oxidation reaction process (Figure 11b). However, the concentration of HC emissions rapidly increased when the air fuel ratio was larger than λ = 1.4 because of the extreme deterioration in the combustion process that resulted from the decreasing combustion temperature and the deterioration of the oxidation reaction of the combustion product. In SI engines, the generation of HC emissions in the combustion chamber is mainly caused by misfiring or incomplete combustion. Such phenomena occur with highly rich or lean air fuel ratios or when the air fuel mixture contains a large amount of burned exhaust or nitrogen that leads to incomplete flame propagation. Additionally, HC emissions are generated by the flame quenching effect, which occurs near the combustion cylinder surface area or from the clearance of deposits of fuel-absorbing oil membranes. In considering the influence of ethanol fuel, the concentration of HC was significantly decreased as the ethanol blending ratio increased. In particular, when the amount of HC emission was compared between E100 and G100, E100 produced approximately 20% less compared to G100 in all excess air fuel ranges. The main reason for decreased concentration of HC emission for blended fuels is that the oxygen in ethanol fuel effectively suppresses misfiring and the formation of incomplete combustion products during the combustion process because of the improved ability of stable and complete combustion. Additionally, ethanol molecules are polar and cannot be easily absorbed by nonpolar molecules in the lubricating oil layer; therefore, ethanol lowers the probability of producing unburned hydrocarbon emissions.19 With regard to the CO emissions of the test fuels, the concentrations rapidly decreased in the range between the stoichiometric ratio (λ = 1.0) and λ = 1.1. In general, the formation of CO during the combustion process is mainly influenced by the excess air ratio. Thus, a large amount of CO is produced at highly fuel-rich conditions because there is a lack of oxygen for the combustion process and carbon contents in the fuel are converted into CO2. Figure 11c demonstrates that as the excess air ratio shifted toward lean combustion, CO emissions sharply decreased. However, CO emissions slightly increased at an excess air ratio of λ = 1.5. This is because an abundance of oxygen sharply reduced CO emissions, while the effect of incomplete fuel burning dissipated when the excess air ratio increased to λ = 1.4. Additionally, the increased concentration of CO emissions at λ = 1.5 resulted from incomplete combustion secondary to the enrichment of oxygen at lean conditions. These results suggest that lean combustion and ethanol fuel can significantly reduce the concentrations of emissions, particularly NOx emission. Figure 12 shows the effect of various excess air ratios and ethanol blend fraction on CO2 emissions under operating conditions of 3000 rpm and 20% engine load conditions. As shown in this figure, it can be seen that the concentrations of CO2 emission for all test fuels are highest at the stoichiometric
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Figure 12. CO2 emission characteristics of test fuels with varying excess air ratios.
air fuel ratio. Because the combustions of test fuels can be achieved the better complete combustion process at this test range. However, as the air fuel excess ratios increased the concentrations of CO2 emissions linearly decreased. Under lean operating conditions, misfiring and partial combustion occur during the combustion process due to the lean effect of the reduction in fuel carbon and the enrichment of air in the combustion chamber. Additionally, as the ethanol fraction increased the CO2 emissions for the blends were lower than that of gasoline fuel, especially in the case of E60. Presumably, the reduced amount of carbon fuel and the increased oxygen content in the fuel resulted in decreased carbon-related emissions from the increasing ethanol fraction and air excess ratio. Therefore, the influence of ethanol at lean combustion conditions can effectively reduce CO2 emission.
4. CONCLUSIONS To investigate the effects of lean combustion with bioethanol gasoline blended fuels, the combustion performance, combustion stability, and reduction characteristics of exhaust emissions were analyzed in a four-cylinder SI engine with varying excess air ratios. The conclusions from this investigation can be summarized as follows: 1 The peak combustion pressures and ROHR of all test fuels linearly decreased as the excess air ratio increased (λ > 1.0). Thus, minimum combustion pressures and ROHR were generated at λ = 1.5. When compared with G100 at given each excess air ratio, there were slight improvements in combustion pressure for the ethanol blended fuels. 2 As ethanol fraction increased, the power and BMEP increased by approximately 1.2% at each air fuel ratio condition compared to G100. Furthermore, the differences in power and BMEP between E100 and G100 were maximized with an increasing air fuel ratio up to λ = 1.5. 3 The peak exhaust gas temperature for all test fuels occurred at λ = 1.0 due to the abundance of air; the gas temperature then linearly decreased with an increasing air fuel ratio. Additionally, the gas temperatures of the ethanol blends were lower than that of gasoline and were generally reduced as the ethanol fraction increased. 4 BSFC increased with the increasing excess air ratio because of lower power output due to the reduced injection mass and incomplete combustion at lean combustion conditions. 3491
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Energy & Fuels In addition, ethanol blends had higher BSFCs compared to G100. The ethanol blends also achieved fairly stable combustion at all excess air ratios compared to gasoline owing to the suppression of unstable combustion and the resulting low levels of COVIMEP. 5 NOx emissions tended to proportionally decrease with increases of the excess air ratio for all test conditions; all of the ethanol blends emitted slightly less NOx compared to G100. The CO2 concentrations for all test fuels were highest at λ = 1.0. However, as the air fuel excess ratios increased the concentrations of CO2 emissions were linearly reduced.
’ AUTHOR INFORMATION Corresponding Author
*Tel: +82-2-2220-0427; fax: +82-2-2281-5286; e-mail: cslee@ hanyang.ac.kr.
’ ACKNOWLEDGMENT This work was supported in part by the Second Brain Korea 21 Project. ’ NOMENCLATURE AFR = Air fuel ratio BMEP = Brake mean effective pressure BTDC = Before top dead center BSFC = Brake specific fuel consumption BSEC = Brake specific energy consumption CO = Carbon monoxide CO2 = Carbon dioxide COV = Coefficient of variation HC = Hydrocarbon LHV = Lower heating value LBL = Lean burn limit MBT = Minimum spark advance for best torque NOx = Nitric oxide ROHR = Rate of heat release RON = Research octane number SI = Spark ignition
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dx.doi.org/10.1021/ef200682b |Energy Fuels 2011, 25, 3484–3492
