Combustion Characteristics of Diesohol Using Biodiesel as an

Publication Date (Web): October 5, 2009 .... Jatropha oil is purchased from a commercial oil shop in Tiruchirappalli, India, and used for ... The expe...
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Energy Fuels 2009, 23, 5413–5422 Published on Web 10/05/2009

: DOI:10.1021/ef900587h

Combustion Characteristics of Diesohol Using Biodiesel as an Additive in a Direct Injection Compression Ignition Engine under Various Compression Ratios V. Arul Mozhi Selvan,* R. B. Anand, and M. Udayakumar Department of Mechanical Engineering, National Institute of Technology, Tiruchirappalli, India Received June 9, 2009. Revised Manuscript Received September 15, 2009

An experimental investigation is carried out to study the combustion characteristics of dieselbiodiesel-ethanol blends in a single-cylinder four stroke direct injection variable compression ratio engine under the compression ratios 15:1, 17:1, and 19:1. As the ethanol is immiscible with diesel, biodiesel (jatropha methyl ester) is used as an additive to prevent the phase separation of the diesel-ethanol blends. The addition of ethanol decreases the cetane number of the blend, whereas the biodiesel addition improves the cetane number of the resultant mixture. The combustion characteristics of the stable fuel blends (D85B10E5, D80B10E10, D75B10E15, D70B10E20, and D65B10E25) are studied and compared with neat diesel by conducting experiments on the computerized variable compression ratio engine test rig. A piezoelectric pressure sensor and a crank angle encoder are used to record the cylinder gas pressure and a crank angle, respectively, to determine the combustion parameters. It is observed that the cylinder gas pressure, maximum rate of pressure rise, and heat release rate increase with higher ethanol concentration due to the longer ignition delay. Also it is found that the ignition delay decreases with the increase in compression ratio and brake mean effective pressure. The total combustion duration is longer as the total fuel consumption is higher for the diesel-biodiesel-ethanol blends than neat diesel. The exhaust gas temperature is found lower for the diesel-biodiesel-ethanol blends. The relative air fuel ratio is found higher for the diesel-biodiesel-ethanol blends than neat diesel and decreases with an increase in brake mean effective pressure. This research work established an insight on the fuel burning characteristics of the diesel-biodiesel-ethanol blends under various compression ratios and loading conditions.

environmental protection and stringent exhaust gas regulations. The use of clean fuel such as natural gas3,4 or the addition of oxygenated fuels such as alcohol, dimethyl ether (DME), etc., to neat diesel provides more oxygen during combustion5 which are ways to reduce the regulated emissions. Fleisch et al.,6 Kapus and Ofner,7 and Sorenson and Mikkelsen8 have studied the utilization of pure oxygenated fuels in diesel engines and found that they can achieve ultralow emissions without a fundamental change in the combustion system. McCormick et al.9 investigated the effect of oxygenate (ethanol and soybean methyl ester) on the regulated emissions. Nord and Haupt10 found that the CO2 reduction is achieved by adding oxygenate to diesel. It is difficult to reduce nitrogen oxides (NOx) and smoke simultaneously in a normal diesel engine because of the trade-off curve between NOx and smoke. A promising method to solve this problem is to provide more oxygen during combustion using an oxygenated alternative fuel such as alcohol.6-8 Ren et al.11 investigated the effect of an oxygenate addition to study the combustion and emission characteristics of a CI engine using diethyl

1. Introduction Combustion characteristics provide valuable information to analyze the engine performance and exhaust emissions which are helpful for engine design and optimization. Studies on combustion characteristics are essential for the optimization of engine design with the use of alternative fuels as compression ignition (CI) engines are basically designed to operate with neat diesel. Even though the literature is replete with studies on diesel-ethanol blends in compression ignition engines, they are mainly concentrated on engine performance and emissions; less work is reported on the combustion parameter analysis based on the heat release process, such as combustion phase analysis and combustion duration analysis.1,2 Besides, the combustion characteristics such as maximum cylinder gas pressure and the heat-release rate can be used to explain the effects of engine operating conditions on the engine performance or to compare the alternative fuels under the same operating conditions. Reduction in engine emissions is a major research aspect in engine development with increasing concern about

(5) Li, W.; Ren, Y.; Wang, X. B.; Miao, H.; Jiang, D. M.; Huang, Z. H. Proc. Inst. Mech. Eng., Part D: J. Automobile Eng. 2008, 222, 265– 274. (6) Fleisch, T.; McCarthy, C.; Basu, A. SAE Trans. 1995, 104 (4), 42– 53. (7) Kapus, P.; Ofner, H. SAE Trans. 1995, 104 (4), 54–69. (8) Sorenson, S. C.; Mikkelsen, S. E. SAE Trans. 1995, 104 (4), 80–90. (9) McCormick, R. L.; Ross, J. D.; Graboski, M. S. Environ. Sci. Technol. 1997, 31 (4), 1144–1150. (10) Nord, K. E.; Haupt, D. Environ. Sci. Technol. 2005, 39 (16), 6260–6265. (11) Ren, Y.; Huang, Z.; Miao, H.; Jiang, D.; Zeng, K.; Liu, B.; Wang, X. Energy Fuels 2007, 21 (3), 1474–1482.

*Corresponding author. Phone: þ919894920835, þ914312503417. E-mail: [email protected]. (1) Murugan, S; Ramaswamy, M. C.; Nagarajan, G. Fuel 2008, 87, 2111–2121. (2) Asfar, K. R.; Hamed, H. Energy Convers. Manage. 1998, 39 (10), 1081–1093. (3) Huang, Z.; Shiga, S.; Ueda, T.; Nakamura, H.; Ishima, T.; Obokata, T.; Tsue, M.; Kono, M. Trans. ASME, J. Gas Turbines Power 2003, 125 (3), 783–790. (4) Shiga, S.; Ozone, S.; Machacon, H. T. C.; Karasawa, T.; Nakamura, H.; Ueda, T.; Jingu, N.; Huang, Z.; Tsue, M.; Kono, M. Combust. Flame 2002, 129 (1-2), 1–10. r 2009 American Chemical Society

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adipate by blending with diesel. The results show that the ignition delay increases with the increase of the diethyl adipate fraction due to the decrease of the cetane number of the blends. Also, the main combustion duration and the amount of heat release in the main combustion duration increase, while the diffusive combustion duration and the amount of heat release in the diffusive combustion duration decrease with the increase of the oxygen mass fraction in the blends. Zhu et al.12 observed longer ignition delay and higher peak pressure and rate of pressure rise while using the oxygenated fuel in a CI engine. The more and faster premixed combustion was due to the longer ignition delay, and the faster diffusion combustion was due to the oxygenated fuel. The composition of the fuel-air mixture influences the rate of combustion and the amount of heat released.13 The higher compression ratio increases the rate of pressure rise and peak temperature inside the cylinder for various fuel blends, hence increase its burning rate. From the heat release rate analysis, Lu14 pointed out that the combustion characteristics of ethanol-diesel blended fuel at higher loads are similar to those of neat diesel due to the cetane number improver (biodiesel) and a large difference exits at lower loads. Czerwinski15 conducted performance tests using a ethanol-rapeseed oil-diesel blend, compared data with neat diesel, and observed longer ignition delay. The addition of ethanol to diesel will increase the ignition delay and combustion noise; therefore, a small addition of a cetane number (CN) improver is helpful to reduce the combustion noise. Diesel-diglyme blends have higher oxygen content and cetane numbers, such as 35.8 and 126, respectively. Ren et al.16 conducted tests using the above blends to study the combustion and emission characteristics of the blends of this nature. The results show that the ignition delay and the amount of heat released in the premixed combustion phase decrease with the increase of the oxygen mass fraction in the blends. The diffusive combustion duration and the total combustion duration decrease, while the amount of heat release in the diffusive combustion phase increases with the increase of the oxygen mass fraction in the blends. Zannis and Hountalas17 observed that the use of oxygenated agents promotes earlier initiation of combustion, and the higher peak pressure is due to the higher cetane number of the blended fuel. Shi et al.18 found that the fuel containing a lower cetane number increases ignition delay and more accumulated fuel/ air mixture, which causes a rapid heat release in the beginning of the combustion, resulting in high temperature and high NOx formation. In the present work, a search for a renewable additive to prevent phase separation between diesel and ethanol blends is performed. After a series of experiments, it is found that jatropha methyl ester prevents phase separation between diesel and ethanol which is nonedible and renewable in nature. Jatropha methyl ester (JME; biodiesel) is produced through a

transesterification process from jatropha curcas oil in the laboratory, and stability studies are performed to identify the minimum quantity of biodiesel (JME) required to keep the diesel-ethanol blends stable. As the cetane number of jatropha methyl ester is higher than those of neat diesel and ethanol, the use of jatropha methyl ester improves the cetane number of the diesel-ethanol blend. The objective of the work is to study the combustion characteristics of various diesel-biodiesel-ethanol blends in a compression ignition engine under various compression ratios to provide valuable information on the use of alternative fuel in a variable compression ratio engine. 2. Experimental Setup and Procedure The stability of ethanol-diesel-biodiesel blends varying in the proportions are investigated at the temperatures of 0, 30, and 45 °C using high speed diesel, ethanol (99.9% purity), and biodiesel. The high speed diesel no. 2 supplied by M/s Indian Oil Corporation Limited through the commercial outlet dealer in Tiruchirappalli, India, is used in this investigation. Ethanol is supplied by M/s Changshu Yangyuan Chemical, China, with 99.9% purity. Jatropha oil is purchased from a commercial oil shop in Tiruchirappalli, India, and used for preparing jatropha methyl ester (JME) through transesterification process in our institute laboratory. The fuel blends are prepared and monitored carefully for stability at 15 min intervals. The experimental setup and procedure followed are described in detail by Arul Mozhi Selvan et al.19,20 It is observed that the stability of the blend is achieved without phase separation for a period of more than a month; hence, the blended fuel can be used reliably for the engine tests. Experiments are conducted in the laboratory to determine the fuel properties of the diesel-biodiesel-ethanol blends. ASTM standards are adopted for testing the fuel properties. The kinematic viscosity of the fuel blends at 40 °C is measured using a viscometer as per ASTM D445 standards. A digital density meter using the ASTM D1298 standard is employed to determine the density of the fuel blends. Flash and fire point testers (ASTM D92) are used to determine the flash and fire points of the diesel-biodiesel-ethanol fuel blends, and an apparatus using the ASTM D93 standard is used for testing neat biodiesel. The ASTM test method D613 is used for the determination of cetane number of the fuel blend. The Pour point of the fuel blends are determined using a pour point apparatus as per ASTM D5985-02 (2008) standard test methods. A copper strip corrosion test is conducted as per ASTM D130 standards. The calorific value of the fuel is determined by bomb calorimeter. The properties of the stable fuel blends are determined in the laboratory, and the values are presented in the Table 1. The stable fuel blends (D85B10E5, D80B10E10, D75B10E15, D70B10E20, and D65B10E25) and neat diesel are used to conduct the experiments in a single cylinder, four stroke, direct injection, variable compression ratio engine to study combustion characteristics under the compression ratios 15:1, 17:1, and 19:1 under steady state conditions at a constant speed of 1500 rpm at various loads. The engine has a provision for changing the compression ratio over a range of 5:1-20:1. The schematic diagram of the experimental setup is shown in Figure 1, and the specification of the engine is shown in Table 2. An eddy current dynamometer is coupled with the engine for precise loading, and

(12) Zhu, R.; Wang, X.; Miao, H.; Huang, Z.; Gao, J.; Jiang, D. Energy Fuels 2009, 23 (1), 286–293. (13) Ramadhas, A. S.; Jayaraj, S.; Muraleedharan, C. Renewable Energy 2006, 31, 1813–1826. (14) Lu, X.-c.; Yang, J.-g.; Zhang, W.-g.; Huang, Z. Energy Fuels 2005, 19 (5), 1879–1888. (15) Czerwinski, J. SAE Trans. 1994, No. 940545. (16) Ren, Y.; Huang, Z.; Miao, H.; Jiang, D.; Zeng, K.; Liu, B.; Wang, X. Energy Fuels 2007, 21 (5), 2573–2583. (17) Theodoros, Z. C.; Hountalas, D. T. Energy Fuels 2004, 18 (3), 659–666. (18) Shi, X.; Yu, Y.; He, H.; Shuai, S.; Wang, J.; Li, R. Fuel 2005, 84, 1543–1549.

(19) Arul Mozhi Selvan, V.; Anand, R. B.; Udayakumar, M. Stability and Performance Characteristics of Diesohol Using Biodiesel as Additive in Compression Ignition Engine. Proceedings of the International conference on Fascinating Advances in Mechanical Engineering, India, Dec 11-13, 2008; pp 667-673. (20) Arul Mozhi Selvan, V.; Anand, R. B.; Udayakumar, M. Stability of Diesohol blend with castor oil as additive and its performance and emission Characteristics in a variable compression ratio engine. Proceedings of the National conference on Energy Security for Rural Development, India, March 26, 2009; pp 110-112.

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Table 1. Properties of Diesel-Ethanol-Biodiesel Fuel Blends properties kinematic viscosity @ 40 °C, cSt density @ 15 °C, gm/cm3 flash point, °C fire point, °C pour point, °C copper strip corrosion cetane number net calorific value, MJ/kg a

diesel a

ethanol

JME

E5

E10

E15

E20

E25

d

5.98 0.893 88 106 -7 1 55.4 38.71

2.91 0.839 17.5 20 -5 1 47.7 40.70

2.67 0.832 15 18 -7 1 46.85 40.50

2.51 0.829 13 17 -10 1 46.1 40.30

2.35 0.827 11 14 -14 1 45.25 40.10

2.2 0.823 9 13 -17 1 44.2 39.90

2 0.83a 50a 56a 6b

1.1314 0.79d 13.5c

46c 42.30

6c 25.18d

-117.3b

Reference 1. b Reference 26. c Reference 18. d Reference 32.

Table 2. Specification of the Engine rated power engine speed compression ratio bore stroke ignition fuel injection timing fuel injection pressure cooling loading system

3.7 kW 1500 rpm (constant) 5:1 to 20:1 (variable) 80 mm 110 mm compression ignition 23° BTDC 20 MPa water cooled eddy current dynamometer

Here, QHR, QHT, P, V, γ, and θ denote heat release energy, heat transfer energy, cylinder gas pressure, cylinder volume, ratio of specific heats, and crank angle, respectively. The first and second term represent the change in sensible energy from the chemical conversion during combustion and associated work imparted on the piston due to the change in pressure. The third term relates the energy loss due to heat transfer effects to the cylinder wall; commonly modeled as a convection term. Integrating the above equation between the ignition crank angle (IGN) and exhaust valve open time (EVO) provides the heat release rate for the entire combustion period.21-23 The determination of the ratio of specific heats is of significant importance for an accurate heat release rate. The parameter’s strong dependency on temperature can be modeled, with considerable accuracy, as a linear relationship:24 γðTÞ ¼ a þ bT ð2Þ where a and b are chosen to be 1.392 and -8.14  10-5 with T in Kelvin.23 The mass fraction burned (MFB) relation comes from a determination of the total possible energy released from the fuel (QCH) as indicated by QCH ¼ mfuel, cycle LHVηc ð3Þ Figure 1. Schematic diagram of the experimental setup.

where, mfuel,cycle, LHV, and ηc represent the mass of fuel, lower heating value of the fuel, and combustion efficiency, respectively. Normalizing the heat release rate (dQHR/dθ) by QCH results in the mass fraction burned curve: R EVO dQHR ð4Þ MFB ¼ IGN QCH

the load applied on the engine is measured using a load cell attached with the arm. A data acquisition system is used to collect, record, and analyze the data from various sensors. A Kistler piezoelectric pressure transducer is used to measure the incylinder gas pressure, a crank angle encoder gives crank angle pulses at every degree crank angle, a K type thermocouple is used to measure the temperature of the exhaust gas, and two infrared sensors mounted at the two ends of a buret with a solenoid valve are used to measure the fuel flow rate. All the signals collected from the sensors are given as input to the data acquisition system through a signal conditioning unit. To eliminate cycle-cycle variation, cylinder pressure data of more than 50 consecutive cycles were collected and averaged using a computer program for proper estimation of cylinder pressure data for heat release analysis as the heat release calculations rely on accuracy the in cylinder pressure data. The following correlations are used to write a computer program to draw the characteristic curves. The heat release of the system is determined by dQHR γ dV 1 dP dQHT ¼ P þ V þ dθ dθ γ -1 dθ γ -1 dθ

In the present investigation, the crank angle between start of injection and 10% mass fraction burned is considered to determine the ignition delay as suggested in the literature. The exhaust gas emissions are measured using the AVL DIGAS analyzer, and the experimental data obtained from the gas analyzer is used for evaluating the relative air-fuel ratio. The estimated uncertainty for the measured and evaluated quantities is presented in Table 3. (21) Krieger, R. B.; Borman, G. L. Trans. ASME 1966, 66, No. WA/ DGP-4. (22) Gatowski, J. A.; Balles, E. N.; Chun, K. N.; Nelson, F. E.; Ekchian, J. A.; Heywood, J. B. SAE Trans. 1984, No. 841359. (23) Cheung, H.; Heywood, J. B. SAE Trans. 1993, No. 932749. (24) Lanzafame, R.; Messina, M. Int. J. Automotive Technol. 2003, 4 (3), 124–133.

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Table 3. Estimated Uncertainty for the Measured and Evaluated Quantities quantity torque (Nm) speed brake mean effective pressure (MPa) fuel injection pressure (MPa) cylinder pressure (bar) fuel injection start angle (deg crank angle) exhaust gas temperature (°C) λ

estimated uncertainty (0.5% (0.5% (1.0% (1.0% (1.0% (0.5 °CA (3 (0.001

3. Results and Discussion The following section 3.1 illustrates the combustion characteristics of the stable diesel-biodiesel-ethanol blends and neat diesel on a variable compression ratio engine. The effect of increasing ethanol content for a constant compression ratio (17:1) and the effect of compression ratio on combustion parameters are discussed in the subsequent sections 3.2 and 3.3, respectively. 3.1. Variation of Combustion Characteristics. The combustion parameters such as the variation of cylinder gas pressure, heat release rate, and mass fraction burned are discussed with reference to the crank angle. The experiments are conducted at various loading conditions, and the economic load is identified at the brake mean effective pressure of 0.44 MPa based on the lowest specific fuel consumption. Hence the above combustion parameters are discussed with the economic loading conditions. The combustion parameters such as peak cylinder gas pressure, rate of pressure rise, maximum heat release rate, ignition delay, combustion duration, relative air-fuel ratio, and exhaust gas temperature are discussed with respect to the brake mean effective pressure. 3.1.1. Variation of Cylinder Gas Pressure. The variation of cylinder gas pressure with crank angle at the brake mean effective pressure of 0.44 MPa for the compression ratios 15:1, 17:1, and 19:1 are shown in Figure 2. It is observed that the peak pressure increases as the compression ratio increases. Also the addition of ethanol and biodiesel accelerates the complete combustion process due to the higher oxygen content of the blended fuel and results higher peak combustion pressure as observed by Zannis and Hountalas.17 At lower compression ratios, the peak cylinder gas pressure for diesel is higher than that for diesel-biodiesel-ethanol blends, and at higher compression ratios, the trend is opposite. This shows that, at higher compression ratios, the diesel-biodieselethanol blends have faster premixed combustion which lead to higher peak cylinder gas pressure. The variation of peak cylinder gas pressure with brake mean effective pressure (bmep) is shown in Figure 3. It is observed that the peak pressure increases with the increase in the brake mean effective pressure and with increase in the compression ratio. Also the peak pressure for the diesel-biodiesel-ethanol blends is higher than that for neat diesel. Longer ignition delay due to the presence of ethanol in the blend would increase the amount of fuel burnt during the premixed combustion phase causing higher peak pressure. The highest cylinder gas pressure obtained is 10.4 MPa at the bmep of 0.55 MPa for the D65B10E25 blend at the compression ratio of 19:1, whereas it is only 8.8 MPa for the same blend at the compression ratio of 15:1. The maximum rate of increase in pressure (MPa/deg) with brake mean effective pressure for various compression ratios

Figure 2. Variation of cylinder pressure with crank angle.

for all the blends is shown in Figure 4. It is observed that the maximum rate of increase in pressure is increasing with increase in the compression ratio as well as with the brake mean effective pressure. The oxygen enrichment in the blend due to the addition of ethanol and biodiesel is the cause for the increased rate of pressure rise.12,17 The highest rate of pressure rise is observed as 0.9570 MPa/deg for the D65B10E25 blend at the compression ratio of 19:1. 3.1.2. Heat Release Rate. The variation of heat release rate with crank angle at the brake mean effective pressure of 0.44 MPa under the compression ratios 15:1, 17:1, and 19:1 are shown in the Figure 5. The heat-release analysis is based on the changes in the cylinder gas pressure and cylinder volume during the cycle. It is observed that the peak heat release rate increases with increases in compression ratio and the addition of up to 20% ethanol and biodiesel in diesel. The use of oxygenated fuels enhances the combustion for all the stable fuel blends and is the cause for the increase in heat release rate.12 Also, the presence of ethanol in the blend increases ignition delay; hence, more fuel is accumulated in 5416

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Figure 4. Variation of max rate of pressure with bmep.

effective pressure. The highest heat release rate is observed as 120 J/deg for the D65B10E25 blend at the bmep of 0.55 MPa under the compression ratio of 19:1, and it is 80.5 J/deg for the same blend and load at the compression ratio of 15:1. 3.1.3. Ignition Delay. Parameters of combustion characteristics such as ignition delay, start of combustion, and the combustion duration are obtained from the heat-release curve. The start of combustion is defined as the point where the heat-release rate turns from negative to zero,25 and the time between the start of injection and the start of detectable heat release is denoted as the ignition delay. The variations of ignition delay with brake mean effective pressure are shown in Figure 7. It is observed that the ignition delay increases with the biodiesel-ethanol addition in the diesel blend and decreases with the increase in compression ratio and brake mean effective pressure. The decrease of the cetane number with a higher ethanol concentration in the diesel-biodiesel-ethanol blend causes longer ignition delay.12 Since the ignition timing delays with

Figure 3. Variation of peak cylinder gas pressure with bmep.

the delay period which results in rapid combustion in the premixed combustion phase causing a higher heat release rate. The heat-release profile has a slight negative dip during the ignition delay period, which is mainly heat loss from the cylinder during the fuel vaporizing phase as reported by Turkcan et al.25 Heat released during the premixed combustion is higher for diesel-biodiesel-ethanol blends than neat diesel due to the longer ignition delay resulting from the lower cetane number and higher latent heat of vaporization of ethanol in the blend.26 The variation of maximum heat release rate with brake mean effective pressure for various compression ratios are shown in Figure 6. It is observed that the maximum heat release rate increases with increase in brake mean (25) Turkcan, A.; Canakci, M. Energy Fuels 2009, 23 (4), 1790–1796. (26) Devan, P. K.; Mahalakshmi, N. V. Appl. Energy 2009, 86, 675– 680.

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Figure 5. Variation of heat release rate with crank angle. Figure 6. Variation of maximum heat release rate with bmep.

the oxygenated fuel addition, more homogeneous air/fuel mixtures were formed during the ignition delay period, and this leads to a rapid premixed combustion14 and is the cause for the in higher in-cylinder gas pressure, higher rate of pressure rise, and higher heat release rate. The biodiesel addition improves the cetane number of the resultant mixture; hence, the ignition delay is relatively shorter for the D85B10E5 blend than D65B10E25. As the cetane number of the ethanol is lower, the higher concentration of ethanol in the diesel-biodiesel-ethanol blend leads to longer ignition delay. The ignition delays for the D65B10E25 blend at the compression ratios of 15:1 and 19:1 are 12.8 and 9.9 °CA, respectively, at the bmep of 0.44 MPa, whereas they are 10.8 and 8.6 °CA for the neat diesel for the above conditions. The ignition delay decreases with higher brake mean effective pressure and compression ratios due to the maximum cylinder gas pressure and the higher temperature with an increasing engine load.25 3.1.4. Combustion Duration. The variations of the total combustion duration with brake mean effective pressure for

various compression ratios are shown in Figure 8. The total combustion duration is the time interval from the start of heat release to the end of heat release. It is observed that the total combustion duration is longer for all the diesel-biodiesel-ethanol fuel blends when comparing them with that of neat diesel. As the calorific value of the diesel-biodiesel-ethanol blend is lower than neat diesel, a greater quantity of fuel is consumed to maintain the engine speed constant at various loads. Hence, the higher total fuel consumption is the cause for longer total combustion duration of the diesel-biodiesel-ethanol blend than that of neat diesel. The premixed combustion is shorter for the dieselbiodesel-ethanol blends than that of neat diesel due to the longer ignition delay with ethanol addition. The diffusive combustion duration of diesel-biodiesel-ethanol blends increases with increase in the engine load. The oxygenated fuel can reduce the over-rich mixture region, improving the mixture quality, and increasing the burning rate; this is more effective for the diffusive combustion phase where oxygen is 5418

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Figure 7. Variation of ignition delay with bmep.

Figure 8. Variation of total combustion duration with bmep.

insufficient. Thus, oxygenated fuel can promote diffusive combustion and reduce soot formation.5,27 At the bmep of 0.44 MPa, the total combustion durations for the fuel blend D65B10E25 and neat diesel are 48.3 and 35.2 at the compression ratio of 15:1, and at the compression ratio of 19:1, they are 43 and 36.9, respectively. 3.1.5. Mass Fraction Burned. The variations of the mass fraction burned with the crank angle for dieselbiodiesel-ethanol blends and neat diesel at the brake mean effective pressure of 0.44 MPa for various compression ratios are shown in Figure 9. As the addition of ethanol increases, the ignition delay and the mass fraction burned are small for diesel-biodiesel-ethanol blends and they decrease with higher ethanol concentrations at a particular crank angle. Hence, more fuel is accumulated in the premixed combustion phase and causes rapid heat release. Due to the higher

oxygen content of the diesel-biodiesel-ethanol blend, the combustion is sustained in the diffusive combustion phase. Also, the higher total fuel consumption of the dieselbiodiesel-ethanol blend causes longer total combustion duration when comparing with neat diesel. 3.1.6. Exhaust Gas Temperature. The variations of the exhaust gas temperature with brake mean effective pressure for various compression ratios are shown in Figure 10. The exhaust gas temperature increases with the higher compression ratio and brake mean effective pressure. The exhaust gas temperatures are lower for the dieselbiodiesel-ethanol blends than those for neat diesel. The highest temperature is recorded as 692 °C for the neat diesel at the bmep of 0.55 MPa at the compression ratio of 19, whereas it is only 616 °C for the D80B10E20 blend at the same bmep under a compression ratio of 15:1. The reason being the lower calorific value of blended fuel as compared to neat diesel and the lower temperature at the end

(27) Banapurmath, N. R.; Tewari, P. G.; Hosmath, R. S. Renewable Energy 2008, 33, 1982–1988.

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Figure 9. Variation of mass fraction burned with bmep.

Figure 10. Variation of exhaust gas temperature with bmep.

of compression. Also, ethanol is partially oxidized (OH radicals), and while burning, it has a lower burning temperature which results in reduced exhaust gas temperatures. These variations in exhaust temperatures can be attributed to the increase in thermal efficiency and/or the decrease in the combustion temperatures.28-30 3.1.7. Relative Air-Fuel Ratio. The variations of the relative air-fuel ratio with brake mean effective pressure for various compression ratios are shown in Figure 11. It is observed that at lower loads, the relative air-fuel ratio is higher and decreases as the load increases. This shows that the engine operates at lean mixtures and approaches toward the stoichiometric region at higher loads. However, it is found that at all the loads the engine operates at lean

mixtures. At higher compression ratios, the mixture becomes relatively richer when compared with lower compression ratios at zero load conditions, whereas at higher load conditions, the variations become very small. At the compression ratios of 15:1 and 19:1 and at the bmep of 0.0 MPa, the relative air-fuel ratios for the fuel blend D65B10E25 are 7.656 and 6.107, whereas at the bmep of 0.44 MPa, they are observed as 2.034 and 1.907, respectively. Similarly for neat diesel, the values are 7.583, 4.948, 1.527, and 1.508 for the above conditions, respectively. 3.2. Effect of Increasing Ethanol Content on Combustion Parameters. The effect of increasing the ethanol content on combustion characteristics for a constant compression ratio is discussed in this section to understand the influence of ethanol percentage on combustion mechanism and performance. The compression ratio of 17:1 is chosen to examine

(28) Ajav, E. A.; Singh, B.; Bhattacharya, T. K. Biomass Bioenergy 1999, 17, 357–365. (29) Agarwal, A. K. Prog. Energy Combust. Sci. 2007, 33, 233–271. (30) Al-Farayedhi, A. A.; Al-Dawood, A. M.; Gandhidasan, P. Trans. ASME, J. Gas Turbine Power 2004, 126, 178–191.

(31) Kwancharareon, P.; Luengnaruemitchai, A.; In, S. J. Fuel 2007, 86, 1053–1061. (32) Lapuerta, M.; Armas, O.; Herrerosm, J. M. Fuel 2008, 87, 25–31.

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release rate is delayed with ethanol concentration. The maximum heat release rate increases with an increase in load and with an increase in ethanol concentration. A 20.3% increase in maximum heat release rate is observed with the use of the E25 blend when compared with neat diesel. The ignition delay increases with increase in the ethanol concentration for the entire load conditions. A 12.9% increase in ignition delay is observed with the use of the E25 blend under the compression ratio of 17:1 and at the bmep of 0.44 MPa. The total combustion duration increases with ethanol concentration and increase in load conditions for all the diesel-biodiesel-ethanol blends. A 26.9% increase in total combustion duration is observed with the use of the E25 blend at the bmep of 0.44 MPa under the compression ratio of 17:1. 3.3. Effect of Compression Ratio on Combustion Parameters. The effect of compression ratio on combustion parameters is discussed in this section by keeping the ethanol blending ratio constant. The influence of variable compression ratio (15:1, 17:1, and 19:1) in the case of E5 and E25 blends are presented to understand the effect of compression ratio on engine combustion parameters with lower and higher ethanol concentrations. At the compression ratio of 15:1, neat diesel produces higher cylinder gas pressure than all the ethanol blends. This is due to the result of incomplete combustion with higher ethanol concentrations at lower compression ratios. Also, it is found that there is a shift on peak pressure with reference to the higher crank angle degrees with an increase in ethanol concentration. This shows that at a lower compression ratio, the ignition delay is longer with more incomplete combustion. At a higher compression ratio of 19:1, the higher ethanol concentration produces a higher cylinder pressure than that for neat diesel. The increase in compression ratio leads to higher cylinder pressure development. In this case, the longer ignition delay with complete combustion produces higher cylinder gas pressure. An increase in 14.7% peak cylinder gas pressure is observed at the compression ratio of 19:1 with the use of the E25 blend when compared with neat diesel at the bmep of 0.44 MPa, whereas at the compression ratio of 15:1, the peak pressure decreases by 9.64% when compared with neat diesel under the same loading condition. The heat release rate increases with an increase in compression ratios. At higher compression ratios of 19:1, the heat release rate of diesel is lower than the ethanol blends. In the case of lower ethanol blends at the compression ratio of 15:1, the occurrence of peak heat release rates for neat diesel is earlier than all the ethanol blends. This is due to the longer ignition delay of the ethanol blends. Due to incomplete combustion of ethanol at a lower compression ratio, the heat release rate is lower for ethanol than neat diesel. The maximum heat release rate increases with an increase in the compression ratio. The ignition delay decreases with increase in the compression ratio and higher load conditions. The total combustion duration decreases with increase in the compression ratio. A 13% reduction in total combustion duration is observed while increasing the compression ratio from 15:1 to 19:1.

Figure 11. Variation of relative air fuel ratio with bmep.

the influence of neat diesel, E5, E10, E15, E20, and E25 blends on engine combustion parameters. The cylinder gas peak pressure is found lower for neat diesel and increases with ethanol concentration. However, the cylinder gas pressure before reaching the peak pressure (earlier crank angle degrees) is lower for higher ethanol concentrations than neat diesel, and after reaching peak pressure, this trend becomes reversed. The addition of ethanol increases the peak pressure up to 5.88% with the E25 blend when comparing with neat diesel. The rate of pressure rise increases with ethanol concentration. The maximum rate of pressure rise increases up to 33% with the use of the blend E25 when comparing with neat diesel at the bmep of 0.44 MPa under the compression ratio of 17:1. The heat release rate increases with ethanol concentration. Heat release rate for neat diesel is lower than all the diesel-biodiesel-ethanol blends. The peak heat release rate shifts toward right (higher crank angle degrees) for the ethanol blends. This shows that the attainment of peak heat

4. Conclusion The combustion characteristics of different diesohol blends are investigated to understand the combustion phenomena of 5421

Energy Fuels 2009, 23, 5413–5422

: DOI:10.1021/ef900587h

Arul Mozhi Selvan et al.

diesel-ethanol blend when biodiesel is used as an additive. The conclusions of this investigation are as follows: 1. The use of diesel-biodiesel-ethanol blends in CI engine causes longer ignition delay which results in higher combustion pressure, maximum rate of pressure rise, and higher heat release rate when compared with the performance of neat diesel. In addition, the peak pressure and rate of pressure rise increases with the increase in brake mean effective pressure and higher compression ratios. The highest peak cylinder gas pressure obtained is 10.4 MPa at the bmep of 0.55 MPa for the D65B10E25 blend at the compression ratio of 19:1, and the highest heat release rate observed is 120 J/deg for the same blend and compression ratio. 2. The ignition delay decreases with higher cetane number of the fuel blend and higher compression ratios. The addition of ethanol in neat diesel decreases the cetane number of the fuel blend; but the biodiesel addition improves the cetane number of the resultant mixture. Hence, the ignition delay is relatively smaller for the blend D85B10E5 than that for E65B10E25. 3. The premixed combustion is faster due to the longer ignition delay and the diffusive combustion is longer as combustion continues due to the use of oxygenated fuel (biodiesel and ethanol) which helps in the reduction of soot. The total combustion duration is longer for all the diesel-biodiesel-ethanol fuel blends when compared with the neat diesel due to the higher total fuel consumption to maintain the engine speed constant at various loading conditions as the calorific value of the diesel-biodiesel-ethanol blend is lesser than neat diesel. 4. The exhaust gas temperature increases with the increase in compression ratio and with brake mean effective pressure and decreases with a higher ethanol proportion in diesel. The highest temperature is recorded as 692 °C for the neat diesel at the bmep of 0.55 MPa at the compression ratio of 19:1, whereas it is only 616 °C for

the D80B10E20 blend at the same bmep under the compression ratio of 15:1. 5. The relative air-fuel ratio is higher at lower bmep values and decreases as the bmep increases. At higher compression ratios, the mixture becomes richer when compared with lower compression ratios. However at all the loads, the engine operates at lean mixtures. Under all the conditions, the diesel-biodiesel-ethanol blend operates at lean mixture than the neat diesel. Acknowledgment. The authors are grateful to Dr. M. Chidambaram, Director, National Institute of Technology, Tiruchirappalli, for granting permission to establish I.C. Engines Research Laboratory in the Mechanical Engineering Department with modern computerized experimental facilities to the international standards. Also special thanks are given to Mr. Palanisamy and Mr. Durairaj for their help rendered during experimentation.

Nomenclature CA = crank angle CR = compression ratio BP = brake power bmep = brake mean effective pressure λ = relative air-fuel ratio CO = carbon monoxide HC = hydrocarbon NO = nitrogen oxide PM = particulate matter JME = jatropha methyl ester D85B10E5 = 85% diesel þ 10% biodiesel þ 5% ethanol (E5) D80B10E10 = 80% diesel þ 10% biodiesel þ 10% ethanol (E10) D75B10E15 = 75% diesel þ 10% biodiesel þ 15% ethanol (E15) D70B10E20 = 70% diesel þ 10% biodiesel þ 20% ethanol (E20) D65B10E25 = 65% diesel þ 10% biodiesel þ 25% ethanol (E25)

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