Article pubs.acs.org/EF
Effect of Bioethanol Blended Diesel Fuel and Engine Load on Spray, Combustion, and Emissions Characteristics in a Compression Ignition Engine Su Han Park,†,⊥ In Mo Youn,‡ Yunsung Lim,§ and Chang Sik Lee*,† †
Department of Mechanical Engineering, Hanyang University, 17 Haengdang-dong, Seongdong-gu, Seoul 133-791, Republic of Korea R&D Planning Team, Korean Institute of Energy Technology Evaluation and Planning, 14 Teheran-ro 114 gil, Gangnam-gu, Seoul 133-502, Republic of Korea § Transportation Pollution Research Center, National Institute of Environmental Research, Environmental Research Complex, Kyongseo-dong, Seo-gu, Incheon 404-708, Republich of Korea ‡
ABSTRACT: This study investigates the effect of engine load condition on the injection spray, combustion, and exhaust emissions characteristics of a diesel−bioethanol blended fuel diesel engine. In this study, the injection characteristics, such as the effective flow diameter and the effective flow velocity, were calculated from the nozzle flow model. The macroscopic spray characteristics of diesel−bioethanol fuels were measured and analyzed using a spray visualization system. Using a four-cylinder test engine with 1.5 L of displacement, the combustion pressure, heat release, and emission characteristics were measured and analyzed. In addition, properties of blended fuel were measured and analyzed. This study revealed that the physical properties (density, viscosity, and surface tension) and chemical properties (cetane number, heating value, and distillation) of diesel− bioethanol blended fuels generally decreased with increased bioethanol content. The increase in bioethanol fuel resulted in easy vaporization at the same temperature condition. After energizing, the increase of engine load caused an increase in spray tip penetration. The spray cone angle was mainly affected by the blending of bioethanol, not the engine load. The increase of engine load led to a decrease of ignition delay by the high gas temperature, and it also caused an increase in the combustion duration for the same fuel amount. In the exhaust emissions, the increase of engine load affected the increase in both the NOx and ISNOx emissions, the decrease of CO and HC emissions. At a high engine load, CO and HC emissions were quite similar in D100, DE10, and DE20 fuels due to the increased oxygen content. reported a number of potential advantages of ethanol use:6 the displacement of imported petroleum with domestic and renewable resources, the significant reduction of diesel particulate matter emissions, the possible improvement in cold flow properties imparted by the ethanol, and the possible improvement in fuel lubricity imparted by emulsifier additives. However, some drawbacks were still reported, such as limited solubility in diesel fuel, weak ignitability, and low dynamic viscosity. Lu et al.7 studied the combustion and emission characteristics of diesel−ethanol blended fuels in a single-cylinder diesel engine. They reported that the diesel−ethanol blended fuels emitted lower carbon monoxide (CO), nitrogen oxides (NOx), and particulate matter (PM) than pure diesel fuel. Also, through the combustion visualization experiment, increases in the ignition delay and shortening of the combustion duration in the diesel−ethanol blended fuels were revealed. Yan et al.8 investigated the combustion and emission characteristics of diesel engines fueled with ethanol−diesel blended fuel in a single cylinder diesel engine. They reported that the increase in ethanol blending ratio caused an extension of the ignition delay, a retardation of the peak combustion pressure, and an increase
1. INTRODUCTION New technologies, such as high-pressure injection and the application of alternative fuels, were developed and have been applied to a diesel engine with high thermal efficiency in order to satisfy strengthened emission regulations. Fossil fuels have been a driving force of industrial development. However, fossil fuel supplies are shrinking due to increased energy consumption. In addition, fossil fuels that have been used in internal combustion engines have now been identified as the main culprit of climate change and environmental pollution. As a result, many countries have gone to great lengths to develop new combustion technology with lower emissions and alternative fuels to replace conventional diesel fuel. Biodiesel, dimethyl ether (DME), bioethanol, and liquefied petroleum gas (LPG) are representative alternative fuels for the internal combustion engine.1−3 Among them, investigations about diesel−ethanol blended fuels are actively progressing. As a fuel for compression ignition (CI) engines, ethanol has certain advantages over diesel fuel, including reduced soot and NOx emissions. Even so, ethanol is currently unable to be used extensively, due to limitations in technology, as well as economic and regional considerations.4 However, ethanolblended diesel fuels can be employed in CI engines.5 Ethanolblended diesel fuel has the added benefit of reducing emissions of diesel particulate matter (PM) from engines employing the fuel. The National Renewable Energy Laboratory (NREL) © 2012 American Chemical Society
Received: May 23, 2012 Revised: July 3, 2012 Published: July 3, 2012 5135
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in the pressure rise rate due to the low cetane number, fast evaporation, and a large latent heat of ethanol fuel. In addition, Yan et al.8 stated that NOx emissions exhibited different results according to the engine load, and that HC and CO emissions increased with higher ethanol blending ratio. Chen et al.9 analyzed the effects of the ethanol blending ratio on spray behavior and exhaust emission characteristics. The spray tip penetration was found to decrease due to the ease of breakup. In addition, the spray cone angle widened, and the PM emissions were reduced by about 20%. Excluding the aforementioned investigations, there have been many studies on the reduction of exhaust emissions and the mixing stability of diesel−ethanol blended fuels.10−12 The purpose of this study is to investigate the effects of engine load on the injection spray, combustion, and exhaust emissions characteristics in a four-cylinder diesel engine using diesel−bioethanol blended fuel. For the analysis of the injection characteristics, the nozzle flow model and injection rate were applied in this study. In addition, the spray tip penetration and the spray cone angle were analyzed. Lastly, the combustion pressure, rate of heat release, cumulative heat release, and exhaust emissions, such as nitrogen oxides (NOx), carbon monoxide (CO), and hydrocarbon (HC), were measured and analyzed for the various engine load conditions and test fuels.
2. EXPERIMENTAL SETUP AND PROCEDURE 2.1. Test Injector and Test Fuels. In this work, a solenoid-type injector with seven holes was used to study the injection, overall spray, combustion, and exhaust emissions characteristics of ethanol-blended diesel fuels. The nozzle hole size, spray angle, and K-factor of the injector were 0.119 mm, 153°, and 1.5, respectively. The K-factor is a design parameter representing the difference between the inlet diameter and outlet diameter of nozzle hole. The operation of the injector was controlled by the injector driver (TEMS, TDA-3200H) in order to change the current signal and energizing duration. The experiments were conducted using three test fuels, including a pure ultra-low-sulfur diesel, and two diesel−bioethanol blended fuels. The bioethanol fuel used in this study had a purity of 99.9% and was anhydrous. According to the literature,13−15 the two major problems associated with ethanol-blended diesel fuels are a low cetane number and phase separation between diesel and ethanol fuels. Therefore, a 10% volume of biodiesel fuel (soybean oil methyl ester) was used to prevent phase separation.15 The physical and the chemical fuel properties such as density, viscosity, heating value, cetane number, surface tension, and distillation curve were measured in order to analyze the effect of a bioethanol fuel on the bioethanol blended diesel fuels. The fuel density and viscosity were measured by the density meter (Anton Paar, DMA 500) and the viscometer (Anton Paar, SVM 3000), respectively. The density meter measured the fuel density of test fuels based on the oscillating U-tube principle, and it had an accuracy of 0.001 g/cm3. The distillation characteristics of test fuels were analyzed by the automatic distillation analyzer (Water Herzog GmbH, HAD 627). The heating value of test fuels was measured by the calorimeter (Parr, Model No. 6400) with an accuracy of 0.1%. The cetane number, which is one of the important factors influencing the combustion engine performance, was measured by the ignition quality tester (Advanced Engine Technology Ltd., IQT). The ignition quality tester was a combustion-based analytical instrument, which enabled the determination of the ignition quality (ignition delay and derived cetane number) of test fuels. 2.2. Injection Rate Measurement System. The injection rates of the test fuels were measured and analyzed using an injection rate measurement system, based on the Bosch method,16 in which the pressure variation in a tube is monitored as fuel is injected into the tube. A piezoelectric pressure sensor (Kistler, 4045A50) was used to measure the pressure variation in the tube. The injection rate meter included an adapter, a measuring tube 5 m in length, a pressure vessel,
Figure 1. Schematic of the experimental apparatuses.
Table 1. Specifications and Setup of the Spray Visualization System item metal-halide lamp
light source power cooling method imaging sensor sensor resolution shutter/shutter speed trigger signal input frame rate
high speed camera
specification metal halide light 150 W air-cooling C-MOS imaging sensor 512 × 512 electronic shutter/1/10 000 s TTL, contact signal 10 000 fps
a relief valve, and a throttle valve. A data acquisition board (National Instrument, NI6013) was used for the acquisition of injection profiles, and the injection quantity was obtained from the mean value of 10 000 continuous injections. In this experiment, the pressure in the tube was set to 3.0 MPa. This system was used to determine the time-resolved injection profile. In this study, the effective flow velocity and the effective diameter were calculated from the injection rate. The effective flow velocity can be obtained from the mean velocity (Umean) and the velocity at the vena contracta, as follows:17,18
Ueff = Uvena −
Umean = 5136
P1 − Pvapor ρl Umean
(1)
ṁ inj ρl Ahole
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2.3. Spray Visualization System and Test Engine. To analyze the overall spray characteristics of bioethanol blended diesel fuels, the spray visualization system were consisted of the spray image acquisition and image processing parts as shown in Figure 1a. Spray images were obtained from a high-speed camera (FASTCAM APX-RS, Photron) equipped with two metal-halide lamps (HVC-SL, Photron) positioned for bottom views. In addition, the high-speed camera was synchronized with an injector signal from the injector driver using a digital delay/pulse generator (Model 555, Berkeley Nucleonics Corp.). Obtained spray images were stored in a computer with an image grabber. A high-pressure chamber, which was capable of pressures up to 4 MPa, was used to investigate the spray characteristics under ambient gas pressure conditions. Ambient gas pressure was controlled using nitrogen gas. The detailed specifications and setup of the spray visualization system are listed in Table 1. To study the effects of engine loads and bioethanol blending ratios on the combustion and exhaust emissions characteristics, a fourcylinder diesel engine with a common-rail injection system and exhaust emissions analyzer was utilized, as shown in Figure 1b. The major specifications of the test engine were a bore of 77.2 mm, a stroke of 84.5 mm, a total displacement of 1.582 L, and a compression ratio of 17.8. A piezoelectric combustion pressure sensor (Kistler, 6055B800) and a data acquisition board were installed to measure the combustion pressure. The common-rail injection system was controlled by an electronic control unit (ECU) equipped with a programmable controller and a crankshaft position sensor to monitor the revolutions per minute (rpm). A universal ECU and an interface system were used to control the fuel injection timing and fuel quantity, respectively. Exhaust emissions were detected and analyzed with a HC, CO, and NOx analyzer (Horiba, Mexa-554JK). The engine loads were varied from 30% to 90%, at 30% intervals. The detailed specifications of the test engine and exhaust emission analyzer are listed in Table 2, and the detailed experimental conditions are listed in Table 3. The reported experimental data for the spray characteristics such as spray tip penetration and spray cone angle were based on an average of five measured data points. The combustion and emissions data were obtained during 300 engine cycles. The measured data, such as combustion pressure and emission data, were averaged.
Table 2. Specifications of Experimental Engine and Exhaust Emission Analyzer experimental engine
exhaust emission analyzer
item
specification
engine type no. cylinders bore × stroke displacement vol. fuel injection system valve type compression ratio engine management system max. power max. torque max. speed model range
4-stroke VGT DI diesel engine 4 77.2 mm × 84.5 mm 1582 cm3 Bosch common rail DOHC 4 valves/cylinder 17.3 Bosch EDC 16
repeatability
response
86 kW @ 4000 rpm 260 N m @ 2000 rpm 4750 rpm MEXA-554JK HC: 0−10 000 ppm vol. CO: 0−10 vol % NOx: 0−4000 ppm HC: ± 12 ppm vol. CO: ± 0.06% vol. NOx: less than ±1.0% HC and CO: 90% response within 10 s NOx: within 30 s
Table 3. Experimental Conditions for Spray Visualization and the Combustion Performance Analysis item test fuels
D100 DE10 DE20
spray visualization
combustion and exhaust emissions
injection pressure (Pinj, MPa) ambient pressure (Pamb, MPa) energizing duration (%) ambient temp. (Tamb, K) fuel temp. (Tfuel, K) engine speed (rpm) injection pressure (Pinj, MPa) start of energizing (deg) engine load (%)
Uvena = Umean/Cc
specification diesel 100% diesel 80% + ethanol 10% + biodiesel 10% diesel 70% + ethanol 20% + biodiesel 10% 70
3. RESULTS AND DISCUSSION 3.1. Physical and Chemical Fuel Properties. The physical properties (density, viscosity, and surface tension) and chemical properties (lower heating value, cetane number, and distillation) of diesel−bioethanol blended fuels are shown in Figure 2. Regarding the fuel density and kinematic viscosity, an increase in either bioethanol content or fuel temperature caused decreases in density and kinematic viscosity. In Figure 2a, the fuel density of DE10 showed the almost similar value with that of D100 because of the addition of biodiesel fuel. The biodiesel fuel, which was added for preventing the phase separation, induced the slight increase of fuel density of bioethanol blended diesel fuel. As shown in Figure 2c, the surface tension exponentially decreased with increased bioethanol content, to a 20% addition, after which a stable tension level was achieved. The lower heating value (LHV) and cetane number of blended fuels decreased as bioethanol content increased because bioethanol fuel has a low cetane number and LHV compared to those of conventional diesel fuel. Figure 2f shows the distillation curve of diesel−bioethanol blended fuels. At the same boiling temperature, the reserved mass of blended fuel was larger than that of the pure diesel fuel. In addition, the reserved mass increased with bioethanol content. This explains why bioethanol fuel is rapidly and easily evaporated in blended fuels, with the evaporation mass increasing with bioethanol content. The active evaporation of bioethanol fuel at low temperature conditions could affect the
1, 3 corresponding conditions to engine loads 30%, 60%, and 90% 290 290 1500 70 BTDC 6 30, 60, 90
(3)
In eqs 1−3, the Ueff, Uvena, Umean mean the effective flow velocity, the velocity at the vena contracta, and the mean velocity, respectively. The P1 and Pvapor indicate the injection pressure and the vapor pressure, respectively. ṁ inf, ρl, and Cc mean the injection rate, fuel density, and the contraction coefficient, respectively. In addition, the effective diameter (Deff) can be calculated from the effective flow velocity and mean velocity, as follows:17,18 Deff = D hole
Umean Ueff
(4) 5137
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Figure 2. Physical and chemical properties of diesel−bioethanol blended fuels.
increase in engine load in each test fuel required increased fuel consumption, which is related to the increase in energizing duration. In addition, for the same engine load conditions with test fuels having a low LHV, the injection quantity should be increased. On the other hand, the right side of Figure 3 shows the effective flow velocity at the nozzle exit. The increase in engine load caused an increase in injection duration. The peak effective flow velocity at the nozzle exit represented the smallest effective diameter (around 0.7 ms after energizing), which can be explained by the inverse relationship of flow area to flow velocity at constant flow quantity. Figures 4 and 5 show the macroscopic spray characteristics, such as the spray development images, spray tip penetration, and spray cone angle, of diesel−bioethanol blended fuels. In Figure 4, the spray images at 0.7 ms (the peak flow velocity condition) and 2.0 ms (condition after the end of injection) after energizing were arranged according to the engine load and test fuels. As shown, the edge of each spray cone had a sharp shape at the low ambient pressure (density) condition (first column in the figure), while a round shape was formed at the
combustion and exhaust emissions characteristics. The results of physical and chemical fuel properties for diesel−bioethanol blended fuels in this investigation showed a similar trend compared to previous literatures.19,20 3.2. Injection and Spray Characteristics. Figure 3 shows the fuel injection characteristics of diesel−bioethanol blended fuels at three engine load conditions. The injection characteristics were analyzed from the effective diameter and effective flow velocity at the nozzle exit. These parameters were obtained from the nozzle flow model suggested by Sarre.17 Detailed information about the model used is described in the research papers of Sarre17 and Park et al.21 In the present study, the injection rate, which was measured by an injection rate meter, was applied to the nozzle flow model. A decrease in the effective diameter denoted the occurrence of cavitation, with an approximate 16.7% decrease observed due to the injector characteristics. Increased engine load caused cavitation to occur sooner and for a longer duration. In addition, an increase in bioethanol content lead to increased cavitation duration. Those results are related to the prolonged injection duration. An 5138
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Figure 3. Injection characteristics of diesel−bioethanol blended fuels according to engine load.
influence on the spray cone angle, although the spray cone angle increased with bioethanol content, as noted in previous studies.22,23 3.3. Combustion and Exhaust Emissions Characteristics. Figure 6 shows the in-cylinder combustion pressure and rate of heat release (ROHR) characteristics of diesel− bioethanol blended fuels according to the engine loads. Test conditions were fixed as 70 MPa injection pressure, 1500 rpm engine speed, and before top dead center (BTDC) 6° injection timing. The engine speed was selected to represent urban driving (1500∼2000 rpm), and the injection timing was determined for ignition around top dead center (TDC). The COVPmax is the coefficient of variance in the maximum incylinder pressure. It can be calculated as the standard deviation divided by the average of the maximum in-cylinder pressure. As shown in Figure 6, the increase in engine load caused an increase in the maximum combustion pressure due to the
high ambient pressure condition (second and third columns in the figure). This demonstrates that the development of the spray tip is easily decelerated in the high ambient pressure. In the energizing period, the spray development of each test fuel for the various engine load conditions showed very similar behaviors. However, after the end of energizing, the increase in engine load caused an increase in spray tip penetration in the three tested fuels. The quantitative analysis of the spray characteristics is illustrated in Figure 6. The spray tip penetration and the spray cone angle at high ambient pressure (Pamb = 3 MPa) are represented in Figure 5. As explained above, variation in engine load has an influence on the spray tip penetration after the completion of the energizing period because the energizing duration is increased under an increased load. The spray cone angle decreased until the peak flow velocity (around 0.7 ms) was achieved, and then it maintained a stable level. The increase in engine load had little 5139
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ratio extended the ignition delay, while little difference in the maximum combustion pressure was observed, except at a 30% engine load. These findings can be explained as the blending of bioethanol fuel with a low cetane number decreased the cetane number of diesel−bioethanol blended fuels and caused an increase in the ignition delay. This phenomenon created sufficient time for the premixing of the injected fuel and ambient gas. Consequently, the sufficiently premixed mixture was rapidly combusted after ignition. Also, the evaporation of bioethanol fuel in blends during the ignition delay affected the formation of a homogeneous mixture. On the other hand, the increase in engine load to 90% induced the increase of the energizing duration because high engine load needed lots of fuel compared to low engine load. With this, the real injection duration was also increased (refer to Figure 3). Due to the long injection duration, the ignition started before the end of injection. Therefore, two peaks, corresponding to the premixed combustion and diffusion combustion were observed in the rate of heat release curve at a high engine load condition. Also, it can be observed that the diffusion combustion region increased.24 In comparison of the 30% and 90% engine loads, the difference in COVPmax remarkably decreased, as quantitatively illustrated in the upper-right side of Figure 6. Figure 7 shows the combustion characteristics such as ignition delay, end of ignition, and combustion duration for the various engine loads and test fuels. The ignition delay was determined as CA10 (the crank angle for 10% cumulative heat release for the total heat release), and the end of ignition was determined as CA90 (the crank angle for 90% cumulative heat release for the total heat release). The combustion duration is the interval between CA10 and CA90. As shown in Figure 7, the ignition delay generally decreased according to increased engine load, while ignition delay increased with the increase in bioethanol content. At the same engine load, the combustion duration decreased with an increase in bioethanol content because the increased ignition delay caused the formation of a sufficient premixed mixture and rapid combustion. Using the same test fuel, an increased engine load caused an increase in combustion duration because the energizing duration also increased due to the need for more fuel. The normalized cumulative heat release characteristics according to the engine load and test fuels are shown in Figure 8. As shown in this figure, the increase in engine load resulted in small ignition delay differences among test fuels. Before ignition, the normalized cumulative heat release was less than zero because of bioethanol fuel evaporation. With an increase in engine load, the portion below the zero level decreased. This is the reason why the maximum cumulative heat release increased with increased engine load, while the reduced heat release according to the evaporation of bioethanol fuel was stable. On the other hand, the increasing gradient of the normalized cumulative heat release became increased for the increase in engine load. This can be explained by the relationship between the real injection duration and ignition timing. The real injection duration of three test fuels was about 6.5° (crank angle) at the low engine load condition (30%), and about 11.4° (crank angle) at the high engine load condition (90%) (refer to Figure 3). Considering the injection timing of BTDC 6 degree and the ignition timing, the combustion at low engine load started after the end of injection. However, the combustion at the high engine load started before the end of injection. In other words, the premixed combustion mainly
Figure 4. Spray images of test fuels at peak injection rate and completion of injection conditions (tasoe: time after the start of energizing (ms). .
increase in injection quantity. Also, a decrease in the ignition delay was observed according to the increase in engine load because the high gas temperature in the combustion chamber promoted ignition and combustion.24 On the other hand, in each engine load condition, the increase in bioethanol blending 5140
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Figure 5. Overall spray characteristics (spray tip penetration and spray cone angle) of test fuels at various engine load conditions. Spray tip penetration and spray cone angle were determined by the average of each spray plume.
known that NOx formation is mainly controlled by the maximum combustion temperature, high temperature duration, and oxygen concentration. The first reason for the reduction in NOx emissions was that the bioethanol blend produced an increased ignition delay and reduced total combustion duration. The second reason was the low LHV of the blended fuel. Therefore, diesel−bioethanol blended fuels require a greater amount of injected fuel for the same engine load conditions compared to those of diesel fuel. This greater amount of fuel correlates to an increased energy requirement for vaporization. Moreover, the higher latent heat of vaporization for bioethanol fuel caused a reduction in the cylinder charge temperature. Consequently, the combustion temperature also decreased with
occurred at the low engine condition. Therefore, the increasing gradient of the normalized cumulative heat release in Figure 8a was steep. On the other hand, in the case of the high engine load with both premixed combustion and diffusion combustion, the increasing gradient had a gentle slope. Figure 9 shows the engine NOx emission characteristics of pure diesel and diesel−bioethanol blended fuels as raw emission (ppm unit) and IS-emission (‘IS’ means the indicated specific g/(kW h) unit). The increase in engine load caused an increase in total NOx raw emission due to the higher injection quantity and combustion temperature.18,25 In the analysis of the raw emission result, the bioethanol blend yielded a slight reduction in NOx emissions at the same engine load. It is well5141
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Figure 6. Combustion and heat release rate characteristics of diesel− bioethanol blended fuels in accordance with the engine load.
a decrease in NOx emission.26,27 On the other hand, in the analysis of the IS-unit emission, regardless of fuel type, the ISNOx emissions of the test fuels showed very similar values at each engine load. This is why the IS-unit denotes exhaust emissions per unit power. Although the test engine was operated at the same engine load conditions for all fuels, the diesel−bioethanol blended fuels had a slightly lower engine power due to the low LHV and the difference in ignition timing compared to those of pure diesel. Therefore, ISNOx emissions were similar for different engine loads, despite the difference in raw NOx emissions. Figure 10 shows the CO and HC emission characteristics of diesel−bioethanol blended fuels at various engine load conditions. Both CO and HC emissions decreased with increased engine load, which indicates that incomplete
Figure 7. Ignition characteristics (ignition delay, end of combustion, and combustion duration) of diesel−bioethanol blended fuels according to engine load at BTDC 6°.
combustion decreased with increased engine load. Actually, the increase in bioethanol content caused incomplete combustion due to the low cetane number and low LHV. The difference in CO and HC emissions between D100 and DE20 decreased with increased engine load. In addition, ISCO and ISHC emissions exponentially increased with the increase in bioethanol content at low engine load condition (30%). However, at moderate and high engine load conditions, the ISCO and ISHC emissions were similar. In other words, if the diesel engine operates in moderate and high load conditions, 5142
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bioethanol fuel disturbed the combustion due to the low cetane number, which was responsible for the sudden increase in ISFC in DE10 and DE20 fuels. However, according to the increase in injection quantity corresponding to engine load, the oxygen component in bioethanol fuel improved the combustion activity and quality. Therefore, at the high engine load condition, the ISFC of the DE20 fuel was only slightly increased compared to that of the D100 fuel.
4. CONCLUSIONS In this study, the effects of engine load variation on the injection, spray, combustion, and exhaust emissions characteristics were investigated in a diesel−bioethanol blended fuel engine. The conclusions derived from various experimental results and parameters are summarized as follows: 1. The physical and chemical properties of diesel− bioethanol blended fuels generally decreased with increased bioethanol content. In addition, the increase in bioethanol fuel content caused the increased vaporization at the same temperature condition. 2. Increased engine load resulted in a longer energizing duration, and the peak effective flow velocity at the nozzle exit was not increased over that of a moderate engine load. On the other hand, a greater injection quantity was required for diesel−bioethanol blended fuel than for pure diesel fuel in order to satisfy the high engine load conditions. Also, bioethanol-blended diesel fuels were consumed at a higher rate at the same engine load, compared to that of pure diesel fuel, due to the low LHV of bioethanol fuel. 3. The engine load affected the spray tip penetration after the completion of energizing and had little influence on the spray cone angle. After energizing, an increased engine load caused an increased spray tip penetration. The spray cone angle was affected by the blending of bioethanol, not by the engine load. 4. Increased engine load led to a decrease in ignition delay due to the high gas temperature in the combustion chamber. The blending of bioethanol fuel caused an increase in ignition delay; however, there was little difference in the maximum combustion pressures moderate and high engine load conditions. At the same engine load, the combustion duration decreased with an increase in bioethanol content. The increase in engine load caused an increase in combustion duration for the same amount of fuel. 5. According to the increase of the engine load, the raw NOx obviously increased, and the ISNOx emission showed the similar values regardless of engine loads. The emission of raw NOx (ppm) decreased by the bioethanol blending effect. However, the ISNOx (g/ kWh) emission maintained a uniform level, regardless of the bioethanol content. On the other hand, the CO and HC emissions decreased with increased engine load and increased with bioethanol content. However, at a high engine load, the emissions (CO and HC) were quite similar in the three test fuels due to the increased oxygen content. 6. The increase in engine load and bioethanol content caused an increase in fuel consumption. According to the increase in bioethanol content, the reduced rate rapidly increased for the increase in engine load.
Figure 8. Normalized cumulative heat release characteristics according to the engine load conditions and test fuels.
the incomplete combustion region (ISHC and ISCO emission region) can be largely reduced. Figure 11 shows the indicated specific fuel consumption (ISFC) characteristics of diesel−bioethanol blended fuels at various engine loads. Generally, the ISFC increased with an increase in bioethanol content. At a low engine load (30%), the bioethanol blending caused an increase in ISFC to about 30.2%. However, at a high engine load condition (90%), the ISFC increased to about 9.6%. In the condition of low engine load, 5143
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Figure 9. Raw NOx and ISNOx emission characteristics of diesel−bioethanol fuels at various engine load conditions.
Figure 11. Fuel consumption characteristics of diesel−bioethanol blended fuels at various engine load conditions.
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ACKNOWLEDGMENTS This work was supported by the Second Brain Korea 21 Project and was supported by the National Research Foundation of Korea (NRF) grant funded by the Korea government (MEST) (No. 2012-0001117)
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Figure 10. CO and HC emissions characteristics of diesel−bioethanol blended fuels at various engine load conditions.
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AUTHOR INFORMATION
Corresponding Author
*Tel: +82-2-2220-0427. Fax: +82-2-2281-5286. Email: cslee@ hanyang.ac.kr. Present Address ⊥
Advanced Photon Source, Argonne National Laboratory, 9700 S Cass Avenue, Lemont, Illinois 60439, United States Notes
The authors declare no competing financial interest. 5144
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dx.doi.org/10.1021/ef300894h | Energy Fuels 2012, 26, 5135−5145
