Effect of Injection Pressure on the Combustion ... - ACS Publications

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Energy & Fuels 2009, 23, 2908–2920

Effect of Injection Pressure on the Combustion, Performance, and Emission Characteristics of a Diesel Engine Fueled with Methanol-blended Diesel Fuel Mustafa Canakci,*,‡ Cenk Sayin,† Ahmet Necati Ozsezen,‡ and Ali Turkcan‡ Department of Mechanical Education, Marmara UniVersity, 34722, Istanbul, Turkey, Department of Mechanical Education, Kocaeli UniVersity, 41380, Izmit, Turkey, and AlternatiVe Fuels R&D Center, Kocaeli UniVersity, 41040, Izmit, Turkey ReceiVed January 20, 2009. ReVised Manuscript ReceiVed April 8, 2009

In this study, the effect of injection pressure on the engine performance, exhaust emissions and combustion characteristics of a single cylinder, four stroke, direct injection, naturally aspirated diesel engine has been experimentally investigated when using methanol-blended diesel fuel from 0 to 15% with an increment of 5%. The engine has original injection pressure of 200 bar. The tests were conducted at three different injection pressures (180, 200, and 220 bar) with decreasing or increasing washer number. All tests were conducted at four different loads (5, 10, 15, and 20 N m) for constant engine speed of 2200 rpm. The experimental test results proved that brake thermal efficiency, heat release rate, peak cylinder pressure, smoke number, carbon monoxide and unburned hydrocarbon emissions reduced as brake-specific fuel consumption, brake specific energy consumption, combustion efficiency, and nitrogen oxides and carbon dioxide emissions increased with increasing amount of methanol in the fuel blend. When comparing the results to the original injection pressure, at the decreased injection pressure (180 bar), peak cylinder pressure, rate of heat release, combustion efficiency, and nitrogen oxides and carbon dioxide emissions decreased, whereas smoke number, unburned hydrocarbon, and carbon monoxide emissions increased at all test conditions. On the other hand, with the increased injection pressure (220 bar), smoke number, unburned hydrocarbon, and carbon monoxide emissions diminished, and peak cylinder pressure, heat release rate, combustion efficiency, and nitrogen oxides and carbon dioxide emissions boosted at all test conditions. With respect to brake-specific fuel consumption, brake-specific energy consumption, and brake thermal efficiency, changing injection pressure gave negative results in the all fuel blends compared to the original injection pressure.

Introduction Recently, the use of diesel engines has increased by virtue of their low fuel consumption and high efficiencies. Nowadays, diesel engines are used in transportation, electric power generation, farming construction, and in many industrial activities. These wide fields of usage lead to increasing requirements of petroleum-derived fuels. The depletion of petroleum reserves and increasing demand also induce a steep rise in fuel prices. It is also known that exhaust emissions from diesel engines cause environmental pollution. Pollutants from diesel engines include carbon monoxide (CO), carbon dioxide (CO2), unburned hydrocarbon (UHC), nitrogen oxides (NOx), and smoke opacities. These emissions have a hazardous effects on air, water, and soil pollution as well as global climatic change and human health. The price of petroleum diesel has soared in recent years, and the available reserves of this important fuel will eventually be exhausted if large-scale use continues; greenhouse gas emission by the usage of fossil fuels is also becoming a greater concern. So, research is now being directed toward the use of alternative renewable and environmentally friendly fuels that are capable of fulfilling an increasing energy demand. In alcohols, methanol and ethanol are used most often as fuels * Corresponding author: Phone:+90 262 3032285; fax: +90 262 3032203; e-mail: [email protected]. † Marmara University. ‡ Kocaeli University.

and fuel additives in diesel engines. Particularly, methanol can be obtained from many sources, both fossil and renewable. These include coal, petroleum, natural gas, biomass, wood, landfills, and even the ocean.1-10 The advantages of methanol as a fuel include:11-13 • High evaporative cooling, which results in a cooler intake process and compression stroke. This raises the volumetric (1) Altun, S.; Bulut, H.; Oner, C. Renewable Energy 2008, 33, 1791– 1795. (2) Saleh, H. E. Fuel 2008, 87, 3031–3039. (3) Icingur, Y.; Hasimoglu, C.; Salman, M. S. Energy ConVers. Manage. 2003, 44, 1745–1753. (4) Wei, Z.; Xu, C.; Li, B. Bioresour. Technol. 2009, 100, 2883–2885. (5) Song, C. L.; Zhou, Y. C.; Huang, R. J.; Wang, Y. Q.; Huang, Q. F.; Lu¨, G.; Liu, K. M. J. Hazard. Mater. 2007, 149, 355–363. (6) Cinar, C.; Topgul, T.; Ciniviz, M.; Hasimoglu, C. Appl. Thermal Eng. 2005, 25, 1854–1862. (7) Thring, R. H. Alternative fuels for spark ignition engines. SAE Paper No. 831685; 1983. (8) Cheng, C. H.; Cheng, C. S.; Chan, T. L.; Lee, S. C.; Yao, C. D. Sci. Total EnViron. 2008, 389, 115–124. (9) Chmielniak, T.; Sciazko, M. Applied Energy 2003, 74, 393–403. (10) Wu, C. W.; Chen, R. H.; Pu, J. Y.; Lin, T. H. Atmos. EnViron. 2004, 38, 7093–7100. (11) Owen, K.; Coley, T. AutomotiVe Fuels Reference Book; Society of Automotive Engineers, Inc.: 1995; pp 386-387, 745-746. (12) Wagner, T. O.; Gray, D. S.; Zarah, B. Y.; Kozinski, A. A. Practicality of alcohols as motor fuel. SAE Paper No.790429; 1979. (13) Adelman, H. Alcohols in diesel engine. SAE Paper No.790956; 1979.

10.1021/ef900060s CCC: $40.75  2009 American Chemical Society Published on Web 05/04/2009

A Diesel Engine Fueled with Methanol-blended Fuel

efficiency of the engine and reduces the required work input in the compression stroke. • Low viscosity compared to diesel fuel, therefore it can easily be injected, atomized, and mixed with air. • Less emission because of its high stoichiometric fuel-air ratio, high oxygen content, high H/C ratio, and low sulfur content. The disadvantages of methanol as a fuel include: • Low energy content of methanol, as seen later in Table 3. This means that almost twice as much methanol as diesel fuel must be burned to give the same energy input to the engine. • Poor cold-weather starting characteristics and danger of storage tank flammability due to low vapor pressure and evaporation. • Poor ignition characteristics due to its low cetane number, high latent heat of vaporization, and high ignition temperature. • More corrosive than diesel fuel on copper, brass, aluminum, rubber, and many plastics. This puts some restrictions on the designs and manifacturing of engines to be used with this fuel. The important advantages and disadvantages of methanol as a fuel for diesel engines are reviewed above. Problems concerning the use of methanol in diesel engines can be different and are briefly described below. Using it in diesel engines as diesel-methanol blends is the simplest method. The most important problem is the phase separation. This problem can be prevented by adding a mixer inside the fuel tank. Also, isopropanol can be added to the methanol-diesel mixtures as an emulsifier to satisfy homogeneity and to prevent phase separation.14 Diesel engine emissions are a highly complex mixture. They consist of a wide range of organic and inorganic compounds distributed among the gaseous and particulate phases. These emissions are a public health concern because the emitted particulates are small (e2.5 µm) and easily respirable and because the particulates contain hundreds of chemicals, some being known carcinogens and mutagens. The emitted gaseous particles contain many irritants and toxic chemicals. Because using methanol-blended diesel fuel can ease off the air pollution, many researchers have been devoted to studying the influence of this alternative fuel on the combustion characteristics, engine performance, and exhaust emissions of IC engines. Song et al.15 investigated the influence of methanol/diesel dual fuel on the performance and emissions of a CI engine. Results showed that BSFC, smoke, and NOx emissions decreased as UHC and CO emissions increased under the dual-fuel operating conditions. Wang et al.16 investigated the effects of methanol mass fraction and pilot diesel injection timing on the ignition to understand the variation in the ignition delay and detailed combustion characteristics. With the increase of methanol mass fraction, the maximum cylinder pressure, the maximum rate of pressure rise, and the maximum heat release rate increased. For emissions, both CO and UHC emissions increased but smoke and NOx emissions decreased. Cheng et al.8 researched the effects of the fumigation methanol on the engine performance, emissions, and particulates. In that study, the fumigation methanol was injected to top up 10, 20, and 30% of the power output under different engine operating conditions. The experimental results showed that there (14) Murayama, T.; Miyamoto, N.; Yamada, T.; Kawashima, J. I. A method to improve the solubility and combustion characteristics of alcoholdiesel blends. SAE Paper No.821113; 1982. (15) Song, R.; Liu, J.; Wang, L.; Liu, S. Energy Fuels 2008, 22, 3883– 3888. (16) Wang, L. J.; Song, R. Z.; Zou, H. B.; Liu, S. H.; Zhou, L. B. Proc. IMechE, Part D 2008, 222, 619–626.

Energy & Fuels, Vol. 23, 2009 2909

is a decrease in the BTE when fumigation methanol is applied, except at the highest load of 0.67 MPa. At low loads, the BTE decreased with the increase in fumigation methanol; but at high loads, it increased with the increase in the fumigation methanol. The fumigation methanol resulted in a significant increase in UHC, CO, and NOx emissions. Yao et al.17 studied the effect of diesel-methanol (DMMC) compound combustion on a diesel engine combustion and emissions. The amount of methanol injected is controlled by an electronic control unit and depends on engine output. Experiments were carried out at idle and five different engine loads at two levels of engine speeds to compare engine emissions resulted from pure diesel and DMMC use, with and without the oxidation catalytic convertor. The results showed that the diesel engine operating with DMMC could simultaneously decrease the soot and NOx emissions but increase the UHC and CO emissions compared with the pure diesel. Udayakumar et al.18 investigated the effect of fumigation methanol on exhaust emissions by using the injection method into the inlet manifold. They carried out their tests with the inlet air heated to 70 °C. They reported that smoke and NOx emissions were both decreased with methanol injection. The effects of methanol-diesel blends on diesel engine emissions were investigated by Ilhan.19 In that study, methanolblended diesel fuels were prepared using 99% pure methanol with the volumetric ratios of 0-15%. The results demonstrated that NOx and CO2 emissions increased as CO and UHC emissions decreased with increasing amount of methanol in the fuel mixture. Chu20 investigated the influence of M0, M5, and M15 methanol/diesel fuel mixture on a single-cylinder diesel engine without changing the engine parameters. Test results showed that methanol addition caused less engine power but improved fuel economy. Although the NOx, smoke, and CO emissions were significantly reduced, UHC emissions increased. Popa et al.21 investigated the effect of the methanol-diesel blend on the exhaust emissions. This paper presented the experimental results obtained by providing two different methods of diesel fuel and methanol engine supplying. The first method consists in the methanol admission through a carburetor combined with the classic diesel fuel injection, and the second one refers to the separate fuels injection. From the exhaust emissions measurements, it was resulted that the smoke and NOx levels significantly reduced for all the engine loads with increasing amount of methanol in the blend. Kulakoglu22 investigated engine performance and exhaust emissions using methanol-diesel blends in a direct injection diesel engine, and the maximum methanol mass fraction was 15%. The results showed that there is an increase in the fuel consumption and NOx emissions and a decrease in the thermal efficiency, CO, and UHC emissions when the methanol amount (17) Yao, C.; Cheung, C. S.; Cheung, C.; Wang, Y.; Chani, T. L.; Lee, S. C. Energy ConVers. Manage. 2008, 49, 1696–1704. (18) Udayakumar, R.; Sundaram, S.; Sivakumar, S. Engine performance and exhaust characteristics of dual fuel operation in DI diesel engine with methanol. SAE paper 2004-01-0096; 2004. (19) Ilhan, M. The Effect of Injection Timing on the Performance and Emissions of a Dual Fuel Diesel Engine; MSc Thesis, Marmara University: Istanbul, 2007; pp 28-30 [in Turkish]. (20) Chu, W. The experimental study about the influence of methanol/ diesel fuel mixture on diesel engine performance. Proceedings of the 2008 Workshop on Power Electronics and Intelligent Transportation System; IEEE Computer Society: WA, 2008; pp 324-327. (21) Popa, M. J.; Negurescu, N.; Pana, C. Resulted obtained by methanol fuelling diesel engine. SAE Paper No.793748; 2001. (22) Kulakoglu, T. Effect of Injection Pressure on the Performance and Emissions of a Diesel Engine Fueled with Methanol-Diesel Blends; MSc Thesis, Marmara University: Istanbul, 2009; pp 43-47 [in Turkish].

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is increased in the mixture. Chao et al.23 studied the effect of a methanol-containing additive on the emissions of carbonyl compounds (CBCs) generated from a heavy-duty diesel engine. When either 10 or 15% methanol-containing additive was used, the emission factors of the CBCs acrolein and isovaleraldehyde increased by at least 91%. In another study,24 the engine performance, combustion characteristics, and the effect of injection timing differences were investigated when an ID diesel engine was operated with diesel fuel and methanol. In that study, diesel fuel was used as a pilot fuel and the ignition lag was mainly controlled by the diesel fuel. The maximum thermal efficiency occurred with the optimum main methanol injection timing. The optimum methanol injection timing was about 2-3° of crank angle before the ignition, as methanol vapor is required for smooth combustion. This was determined from results showing that the methanol combustion was initiated by flame propagation from the combustion of diesel fuel. Fuel injection pressure is a significant operating parameter affecting the performance and emissions in CI engine. When fuel injection pressure is low, fuel particle diameters will enlarge and ignition delay period during the combustion will increase. CO, UHC, and NOx emissions will increase since the combustion process goes to a bad condition. When injection pressure is increased, fuel particle diameters will become small. Since formation of mixing of fuel to air becomes better during the ignition period, smoke level and CO emissions will be less. But, if injection pressure is too high, the ignition delay period becomes shorter. So, possibilities of the homogeneous mixing decrease. Therefore, smoke is seen at the exhaust of the engine.25,26 Various investigators have reported that injection pressure has an effect on exhaust emissions. Can et al.27 studied the effects of ethanol addition (10 and 15% volume) to diesel fuel on the emissions of a diesel engine having different injection pressures (150, 200, and 250 bar) at full load. Experimental results showed that increasing the injection pressure of the engine running with ethanol-diesel decreased CO and smoke emissions while it increased NOx emissions. Celikten26 investigated the influence of injection pressure on exhaust emissions of a diesel engine. In this study, emission values had been measured with both full and part loads by changing the injection pressure from 100 to 250 bar with the interval of 50 bar and for throttle positions of 50, 75, and 100%. According to the test results, high injection pressure for O2, SO2, and CO2, low injection pressure for NOx and smoke level must be preferred for decreasing emissions. Sekmen et al.28 researched the smoke level of a diesel engine at five different injection pressures (100, 150, 175, 225, and 225 bar). Their test results showed that increasing the injection pressure reduced smoke level, and increasing the injection pressure boosted smoke level. Icingur and Altiparmak29 examined the effect of injection pressures and fuel cetane numbers on a diesel engine. For this (23) Chao, H. R.; LI˙N, T. C.; Chao, M. R.; Chang, F. H.; Huang, C. I.; Chen, C. B. J. Hazard. Mater. 2000, B73, 39–54. (24) Enoki, K.; Hayashi, S.; Sawa, N. Optimum injection timings of gas-oil and methanol in dual fuel CI engine. SAE Paper No. 932479; 1993. (25) Bakar, R. A.; Ismail, S.; Ismail, A. R. Am. J. Appl. Sci. 2008, 5 (3), 197–202. (26) Celikten, I. Appl. Thermal Eng. 2003, 23, 2951–2060. (27) Can, O.; Celikten, I.; Usta, N. Energy ConVers. Manage. 2004, 45, 2429–2440. (28) Sekmen, Y.; Cınar, C.; Erduranlı, P.; Boran, E. J. Polytechnic 2004, 7 (4), 321–326, In Turkish. (29) Icingur, Y.; Altiparmak, D. Turk. J. Eng. EnViron. Sci. 2003, 27, 291–297.

Canakci et al.

purpose, fuels with 46, 51, 54.5, and 61.5 cetane numbers were tested in a diesel engine at four different injection pressures (100, 150, 200, and 250 bar). The results indicated that increasing the injection pressure reduced smoke level and increased NOx and SO2 emissions. Cinar et al.30 investigated the effect of injection pressure and intake CO2 concentration on exhaust emissions. The investigation was carried out in a direct injection and turbo-charged diesel engine. It was concerned with the effect of using diluting CO2 in the intake manifold and injection pressure on BSFC and smoke level and CO emissions. The tests demonstrated that BSFC deteriorated with increasing injection pressure and CO2 concentration. On the other hand, smoke level and CO emission proportionally deteriorated with increasing CO2 concentration, and injection pressure did not affect smoke level and CO emission. As seen in the literature review, it can be realized that methanol-blended diesel fuel can effectively reduce the pollutant emissions. On the other hand, the influence of injection pressure on the exhaust emissions has not been clearly studied when using methanol-blended diesel fuel in the CI engines. Therefore, in this study, both the effects of injection pressure and methanolblended diesel fuel on the engine performance and exhaust emissions were experimentally investigated using a singlecylinder direct injection diesel engine. Experimental Apparatus and Measurement The experimental set up is shown in Figure 1. The engine used in this study is a single-cylinder, naturally aspirated, air-cooled, direct injection diesel engine. The basic data of the engine are given in Table 1. A Cussons-P8800 type standard engine test equipment consists of electrical dynamometer and measurement instruments used in the experiments. The dynamometer is a DC machine rated at 10 kW and 380 V. The load on the dynamometer was measured using a strain gauge load sensor. An inductive pickup speed sensor was used to measure the speed of the engine. The pressure-time history of the cylinder was measured using a Kistler Model 6052B air-cooled piezo-quartz pressure sensor, which was mounted on the cylinder head. The signals were then passed onto a Kistler Model 5644A charge amplifier. Crankshaft position was obtained using a crankshaft angle sensor to determine cylinder pressure as a function of crank angle. The crank angle signal was obtained from an anglegenerating device mounted on the main shaft. The signal of cylinder pressure was acquired for every 0.75 °CA, and the acquisition process covered 100 completed cycles. Coolant temperature, exhaust temperature, inlet air temperature, and engine oil temperature were measured using K type thermocouples. CO, CO2, and UHC emissions were measured with an infrared gas analyzer (Bilsa Mod 210) with an accuracy (0.01%, (0.01%, and (1 ppm, respectively. Smoke levels were obtained using a Bosch system with an accuracy (0.1%. NOx emissions were recorded using an electrochemical gas analyzer (Kane-May Qintox KM9106) with an accuracy of (1 ppm. The emission data were expressed as “brake-specific” basis (g/kWh) except for the Bosch smoke number. Brake-specific emissions are the mass flow rate of the pollutant divided by the engine power. Fuel consumption was quantified by the combined container method. Pressure in the intake manifold was determined by using an inclined manometer. The accuracies of the measurements and the uncertainties in the calculated results are given in Table 2. To prepare the methanol-blended fuel mixture, two fuels (diesel and methanol) were used. The conventional diesel fuel was obtained from the TUPRAS Petroleum Corporation. Methanol, with a purity of 99%, was purchased from a commercial supplier. The properties of the diesel fuel and methanol are summarized in Table 3. The methanol was blended with Euro diesel fuel at the ratios of 0/100, (30) Cinar, C.; Topgul, T:; Ciniviz, M.; Hasimoglu, C. Appl. Thermal Eng. 2005, 25, 1854–1862.



Figure 1. Experimental setup. Table 1. Technical Specifications of the Test Engine47 engine type cylinder number bore stroke total cylinder volume injector opening pressure number of nozzle hole start of injection timing compression ratio maximum torque maximum power

Lombardini 6 LD 400 1 86 mm 68 mm 395 cm3 200 bar 4 20 °CA BTDC 18:1 21 N m at 2200 rpm 7.5 kW at 3600 rpm

Table 2. The Accuracies of the Measurements and the Uncertainties in the Calculated Results measurements

accuracy

Load Speed Time Temperatures

(2 N m (25 rpm (0.5% (1 °C

calculated results

uncertainty

power BSFC BSEC

(2.55% (2.60% (2.60%

Table 3. Properties of the Fuels Used in the Tests formula molecular weight (kg/kmol) boiling temperature (°C) density (g/cm3, at 20 °C) flash point (°C) autoignition temperature (°C) lower heating value (MJ/kg) cetane number viscosity at 298.15 K (mPa s) stoichiometric air-fuel ratio heat of vaporization (MJ/kg)

methanol45

Euro diesel46

CH3OH 32 64.7 0.79 11 470 20.27 4 0.59 6.66 1.11

C14.34H24.75 196.8 190 - 280 0.83 78 235 42.74 56.5 3.35 14.28 0.27

5/95, 10/90, and 15/85 (by vol). The fuel blends were prepared just before starting the experiment to ensure that the fuel mixture is homogeneous. A mixer was also used in the fuel tank in order to prevent phase separation.

Prior to starting the engine, the nozzle was taken off and adjusted to 200 bar, which is factory demand. The washer, located in the connection place between nozzle and injector spring, is 0.20 mm, and adding one washer increases the injection pressure by 20 bar. Experiments were carried out at three different injection pressure (180, 200, and 220 bar) values with decreasing or increasing washer number. All tests were conducted at four different loads (5, 10, 15, and 20 N m) at the constant engine speed of 2200 rpm. The values of engine oil temperature, mass flow rate of air, exhaust temperature, and pollutants such as SN, CO, CO2, UHC, and NOx were recorded during the experiments. All data were collected after the engine stabilized. Estimation of the Experimental Heat Release Rate and Combustion Efficiency. Heat release analysis can yield valuable information about the effect of engine design, fuel injection system, fuel type, and engine operating conditions on the combustion process and engine performance.31 In this study, the cylinder pressure data were used to evaluate the rate of heat release (ROHR), which is a simplified thermodynamic model. The ROHR was calculated using the first-law analysis of thermodynamics. The ROHR at each crank angle was determined by the following formula:

Q)

1 γ (P dV) + (V dP) + Qw γ-1 γ-1

(1)

where Q is the apparent heat release rate (J), γ is the ratio of specific heats that is calculated according to an empirical equation,32 P is the cylinder pressure (Pa), V is the instantaneous volume of the cylinder (m3), and Qw is heat transfer rate (J) from the wall calculated based on the Hohenberg correlation,33 and the wall temperature was assumed to be 723 K. For this calculation, the contents of the cylinder were assumed to behave as an ideal gas (air) with the specific heat being dependent on temperature; leakage through the piston rings was neglected.34 (31) Ghojel, J.; Honnery, D. Appl. Thermal Eng. 2005, 25, 2072–2085. (32) Brunt, M. F. J.; Rai, H.; Emtage, A. L. The calculation of heat release energy from cylinder pressure data. SAE Paper No. 981052; 1998. (33) Hohenberg, G. H. M. Advanced approaches for heat transfer calculations. SAE Paper No. 790825; 1979. (34) Hayes, T. K., Savage, L. D.; Sorenson, S. C. Cylinder pressure data acquisition and heat release analysis on a personal computer. SAE Paper No. 860029; 1986.

98.45 (-) 98.56 (0.11) 98.77 (0.32) 98.97 (0.52) 98.32 (-) 98.46 (0.13) 98.52 (0.20) 98.66 (0.34) 98.33 (-) 98.46 (0.13) 98.49 (0.16) 98.53 (0.20) 98.13 (-) 98.14 (0.01) 98.27 (0.14) 98.41 (0.28) 16.5 (-) 16.25 (-1.51) 15.22 (-7.75) 13.73 (-16.76) 15.42 (-) 15.32 (-0.62) 15.03 (-2.55) 14.71 (-4.59) 270.2 (-) 278.3 (3) 283.8 (5.04) 303.2 (12.21) M0 M5 M10 M15

150.8 (-) 155.9 (3.39) 163.5 (8.39) 171.8 (13.92)

144.1 (-) 151.4 (5.03) 159.4 (10.63) 171.8 (19.22)

141 -) 147.1 (4.32) 161.4 (14.5) 184.1 (30.57)

11.62 (-) 11.65 (0.25) 11.55 (-0.56) 11.99 (3.23)

6.49 (-) 6.53 (0.62) 6.65 (2.62) 6.8 (4.81)

6.2 (-) 6.33 (2.23) 6.49 (4.74) 6.8 (9.68)

8.61 (-) 8.59 (-0.25) 8.65 (0.56) 8.34 (-3.13) 180 bar 6.06 (-) 6.15 (1.53) 6.57 (8.4) 7.28 (20.13)

16.14 (-) 15.79 (-2.18) 15.41 (-4.52) 14.71 (-8.83)

99.16 (-) 99.37 (0.21) 99.39 0.23 99.79 (0.63) 99.10 (-) 99.24 (0.14) 99.31 (0.21) 99.58 (0.48) 98.99 (-) 99.12 (0.13) 99.15 (0.16) 99.29 (0.30) 98.89 (-) 99.01 (0.12) 99.04 (0.15) 99.18 (0.29) 21.16 (-) 20.31 (-4.01) 20.16 (-4.7) 18.44 (-12.84) 17.57 (-) 17.54 (-0.18) 17.31 (-1.49) 16.87 (-3.99) 259.4 (-) 264.1 (1.82) 271.9 (4.81) 286.3 (10.39) M0 M5 M10 M15

132.3 (-) 136.2 (2.93) 141.9 (7.23) 149.8 (13.21)

127.2 (-) 130.7 (2.81) 132.4 (4.16) 145.6 (14.5)

109.9 (-) 117.7 (7.04) 121.8 (10.84) 137.1 (24.72)

11.15 (-) 11.05 (-0.9) 11.07 (-0.77) 11.33 (1.56)

5.69 (-) 5.7 (0.18) 5.78 (1.51) 5.93 (4.16)

200 bar (ORG Injection Pressure) 5.47 (-) 4.73 (-) 8.97 (-) 5.47 (0.06) 4.92 (4.18) 9.05 (0.91) 5.39 (-1.39) 4.96 (4.94) 9.04 (0.78) 5.76 (5.34) 5.42 (14.74) 8.83 (-1.53)

18.29 (-) 18.28 (-0.06) 18.55 (1.41) 17.36 (-5.07)

99.18 (-) 99.33 (0.15) 99.56 (0.38) 99.71 (0.53) 99.09 (-) 99.23 (0.14) 99.37 (0.28) 99.66 (0.57) 99.06 (-) 99.20 (0.14) 99.30 (0.24) 99.49 (0.43) 18.29 (-) 17.36 (-5.1) 17.5 (-4.32) 15.69 (-14.19) 16.14 (-) 16.06 (-0.46) 16.36 (1.37) 16.09 (-0.32) 264.7 (-) 269.6 (1.86) 277.7 (4.92) 292.9 (10.64)

144.1 (-) 148.8 (3.22) 150.2 (4.2) 157.1 (9.04)

135.1 (-) 139.2 (3) 143.5 (6.2) 159.1 (17.75)

127.2 (-) 137.7 (8.28) 140.4 (10.39) 161.1 (26.67)

11.38 (-) 11.28 (-0.86) 11.31 (-0.67) 11.59 (1.79)

6.2 (-) 6.23 (0.46) 6.11 (-1.35) 6.22 (0.32)

5.81 (-) 5.82 (0.25) 5.84 (0.55) 6.29 (8.33)

8.79 (-) 8.86 (0.87) 8.84 (0.67) 8.63 (-1.76)

17.21 (-) 17.17 (-0.25) 17.12 (-0.55) 15.89 (-7.69)

15 N m 10 N m 5Nm 20 N m 15 N m

220 bar 5.47 (-) 5.76 (5.38) 5.71 (4.51) 6.37 (16.54)

10 N m 5Nm 20 N m 15 N m 5Nm 20 N m 15 N m

M0 M5 M10 M15

(35) Canakci, M. Bioresour. Technol. 2007, 98, 1167–1175. (36) Korkmaz, I. A Study on the Performance and Emission Characteristics of Gasoline and Methanol Fuelled Spark-ignition Engines; Ph.D. Thesis, Istanbul Technical University: Turkey, 1996; p 49.

10 N m

Diesel engine exhaust emissions have the potential to cause a range of health problems. The application of methanol as a supplementary CI engine fuel may reduce exhaust emissions. However, to reach the emission reduction, it may require some modification on the engine. The injection pressure in diesel engines plays an important role for emission control strategies. So, the influence of injection pressure and methanol-blended diesel fuel on the combustion characteristics and exhaust emissions of a single-cylinder CI engine was experimentally investigated. The experimental conditions were selected as follows: four engine loads (5, 10, 15, and 20 N m), 2200 rpm constant speed, and three injection pressures (180, 200, and 220 bar). The fuels were M0, M5, M10, and M15, indicating the content of methanol in different volume ratios (e.g., M5 contains 5% methanol and 95% diesel fuel by volume). Engine Performance. Table 4 shows the BSFC, BSEC, BTE, combustion efficiency values, and the percent changes in these parameters compared to M0 (diesel fuel), respectively, at different engine loads and injection pressures. Brake-specific Fuel Consumption (BSFC). The BSFC is described as the ratio of the fuel consumption to the brake power. As shown in Table 4, the change in BSFC relative to M0 is 2.93, 7.23, and 13.21% for M5, M10, and M15, respectively, at 10 N m and ORG injection pressure (200 bar). The results show that increasing methanol ratio in the fuel mixture leads to increase in BSFC. This behavior is attributed to the LHV of methanol, which is distinctly lower than that of the diesel fuel.36 Therefore, the amount of fuel introduced to the cylinder for a desired energy input has to be greater with the methanol fuel. The BSFC decreases about 1.85 times as the engine load increases from 5 to 20 N m constant loads

5Nm

Result and Discussions

fuel type

where ηcombustion is the combustion efficiency, LHVfuel is the lower heating value of the fuel (kJ/kg-fuel), HNO is ht enthalpy of formation of NO (kJ/g-NO), HNO2 is the enthalpy of formation of NO2 (kJ/g-NO2), HCO is the enthalpy of formation of CO (kJ/gCO), NO is the exhaust emission level of NO (g/kg-fuel), NO2 is the exhaust emission level of NO2 (g/kg-fuel), CO is the exhaust emission level of CO (g/kg-fuel), and UHC is the exhaust emission level of UHC (kg/kg-fuel).

Table 4. BSFC, BSEC, Thermal and Combustion Efficiencies of the Engine

(2)

10 N m

× 100 brake thermal efficiency (% change)

LHVfuel

BSEC, MJ/kWh (% Change)

ηcombustion ) LHVfuel - HNONO - HNO2NO2 - HCOCO - LHVfuelUHC

combustion efficiency (% change)

Because the fuel’s chemical energy is not fully released inside the engine during the combustion process, it is useful to define combustion efficiency. Thus, the combustion efficiency was calculated for the tested fuels in the engine. The combustion energy losses take into account the energy required to form NO and NO2 and the energy lost owing to incomplete oxidation of CO to CO2 and UHC fuel to CO2 and H2O. The enthalpy of formation for H2O and CO2 from fuel oxidation was not calculated theoretically, instead the LHV was used. Each enthalpy of formation calculation is mass-specific to the emission quantity measured during the engine testing.35 The enthalpy of formation for the measured exhaust emissions is used to calculate the combustion efficiency in accordance with the equation below.

99.26 (-) 99.43 (0.17) 99.68 (0.42) 99.90 (0.64)

Canakci et al. 20 N m


BSFC, g/kWh (% change)

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for M15. This decrease in BSFC could be explained by the fact that as engine load increased, the rate of increasing brake power was much more than that of the fuel consumption. As shown in Table 4, the minimum BSFC values were attained at ORG injection pressure for the all fuel blends. When the injection pressure was increased (220 bar) and decreased (180 bar) compared with ORG injection pressure, the change in BSFC was measured as 6.5 and 15.8%, respectively, for M5 at 15 N m. With decreasing injection pressure, fuel particle diameters will enlarge and ignition delay period during the combustion will increase. This situation causes an increase in the BSFC. On the other hand, increasing injection pressure causes a shorter ignition delay period. So, possibilities of homogeneous mixing decrease and BSFC augments.25 Brake Specific Energy Consumption (BSEC). The BSEC is defined as multiplication of BSFC and LHV. As shown in Table 4, the BSEC increases with methanol content. The minimum BSEC obtained was 4.73 MJ/kWh for M0, 4.92 MJ/kWh for M5, 4.96 MJ/kWh for M10, and 5.42 MJ/kWh for M15 at 20 N m load and ORG injection pressure, respectively. It is wellknown that the LHV of the fuel affects the engine power. The lower heat content of the methanol-diesel blend causes some reductions in the engine power. In addition, the theoretical air-fuel ratio of diesel fuel is about two times higher than that of methanol, as shown in Table 3. For these reasons, the effective power should decrease with the increase of methanol amount in the fuel mixture. Thus, the engine needs more heat consumption to maintain the same amount of power output.37,38 The BSEC decreases with increasing engine loads due to noticeably diminishing BSFC for the all fuel blends and injection pressure. The BSEC reduces by 43% as the engine load increases from 5 to 20 N m constant loads for M10 at decreased injection pressure. When the injection pressure changed from ORG injection pressure, BSEC values increased due to the increase in the energy requirement to sustain the same amount power output at ORG injection pressure. The increments for the increased and the decreased injection pressures were 13.24 and 6.25% for M0 at 15 N m, respectively. Brake Thermal Efficiency (BTE). The BTE is described as the ratio of the brake power to fuel consumption and LHV. As shown in the Table 4, compared to M0, the BTE diminished by 4, 4.7, and 12.8% for M5, M10, and M15, respectively, at 20 N m and ORG injection pressure. BTE points to the ability of the combustion system to accept the experimental fuel and provides comparable means of assessing how efficient the energy in the fuel was converted to mechanical output. From the previous discussion, it could be concluded that as the methanol amount increases in the fuel blend, the BSFC increases, since the LHV value of the blend decreases. As mentioned above, BTE is a function of BSFC and LHV of the blend for a constant effective power. It is clear that BSFC is more effective than LHV with regard to increasing BTE. Therefore, the BTE increased as the methanol content decreased in the blended fuel for all injection pressures.39 Increasing engine loads caused an increase in BTE values owing to a noticeable decline in BSFC for the all fuel blends and injection pressures. The BTE increased by 13.3% as the engine load increased from 10 to 20 N m at constant loads for M0 at increased injection pressure. The best results in terms of BTE were obtained at ORG injection pressure. Decreased or increased injection pressure decreased BTE values by increasing

BSFC. When the injection pressure was increased and decreased in comparison to ORG injection pressure, BTE decreased by 13.5 and 22% for M0 at 20 N m load, respectively. Combustion Efficiency. The change in combustion efficiency increased with increasing methanol ratio in the fuel blend at the all engine loads and injection pressures compared to M0. As demonstrated in Table 4, the change in the combustion efficiency increased by 0.14, 0.24, and 0.43% for M5, M10, and M15, respectively, at 5 N m engine load and increased injection pressure relative to M0. When methanol is added into diesel fuel, the fuel contains more oxygen, which reduces CO and UHC emissions and increases NOx emissions. These effects caused an increase in combustion efficiency as shown in eq 2. Combustion efficiency slightly augmented with increasing engine load from 5 to 20 N m for all test fuels due to better volumetric efficiency and atomization rate. The maximum combustion efficiency (99.9%) was obtained with increased injection pressure, followed by ORG injection pressure (99.79%) and decreased injection pressure (98.97%) for M15 at 20 N m load. The increase in combustion efficiency with the increased injection pressure was attributed to increases in the NOx emissions and decreases in CO and UHC emissions, as mentioned below. Exhaust Emissions. The exhaust emissions measured were UHC, NOx, smoke number (SN), CO, and CO2. Table 5 and Figures 2-5 demonstrate the emission values and percent changes in the emissions at different engine loads and injection pressures compared to M0, respectively. Unburned Hydrocarbon (UHC) Emissions. Most of the UHC is caused by an unburned fuel-air mixture, whereas the other source is the engine lubricant and incomplete combustion. The term UHC means organic compounds in the gaseous state; solid HCs are the part of the particulate matter. Typically, HCs are a serious problem at low loads in CI engines. At low loads, the fuel is less apt to impinge on surfaces; but, because of poor fuel distribution, large amounts of excess air and low exhaust temperature, lean fuel-air mixture regions may survive to escape into the exhaust.40 Relating to the effect of different methanol ratios on UHC emission, it was found that increasing methanol ratio in the fuel blend decreased UHC emissions. For example, the UHC emissions compared to M0 at ORG injection pressure and 20 N m load decreased by 18, 33, and 46% for M5, M10, and M15, respectively, as seen in Figure 5b. When methanol was added to the diesel fuel, it provided more oxygen for the combustion process and led to the improving combustion. In addition, methanol molecules are polar and cannot be absorbed easily by the nonpolar lubrication oil; and therefore methanol can lower the possibility of the production of UHC emissions.41 Increased engine load decreased UHC emissions that were in the same trend with CO. For example, comparing Figure 2c to Figure 5c, it was observed that the change in the UHC emissions diminished by 14% for M10 as the engine load increased from 5 to 20 N m constant loads at decreased injection pressure. Figure 4, panels a-c, illustrates the change in UHC emissions with different methanol blends at different injection pressure for 15 N m load compared to M0. From these figures, it was found that increased injection pressure caused a reduction in UHC emission by 13.6% and decreased injection pressure increased the UHC emission by 7% compared to ORG pressure for M15, respectively. The increasing injection pressure caused

(37) Sayin, C.; Uslu, K.; Canakci, M. Renewable Energy 2008, 33, 1314– 1323. (38) Rakopolous, C. D.; Kyristis, D. C. Energy 2001, 26, 705–722. (39) Sayin, C.; Uslu, K. Int. J. Energy Res. 2008, 32, 1006–1015.

(40) Sayin, C.; Ertunc, H. M.; Hosoz, M.; Kilicaslan, I.; Canakci, M. Appl. Thermal Eng. 2007, 27, 46–54. (41) Alla, G. H.; Soliman, H. A.; Badr, O. H.; Rabbo, M. F. Energy ConVers. Manage. 2002, 43, 269–277.

0.99 0.93 0.90 0.81 0.86 0.81 0.80 0.71 0.76 0.74 0.74 0.66 0.72 0.71 0.70 0.63 33.03 36.13 39.16 49.63 25.22 26.11 27.42 34.11 15.01 14.88 18.04 20.14 17.44 25.14 27.16 34.96 0.41 0.33 0.26 0.22 0.61 0.49 0.44 0.40 180 bar 0.71 0.64 0.60 0.53 1.17 0.99 0.94 0.82 1.69 1.82 2.01 2.54 1.53 1.66 1.85 2.26 1.20 1.30 1.43 1.79 1.18 1.26 1.37 1.79 0.41 0.35 0.29 0.28 0.45 0.38 0.38 0.31

0.83 0.73 0.68 0.59

1.09 0.96 0.93 0.80

M0 M5 M10 M15

M0 M5 M10 M15

0.55 0.44 0.39 0.35

0.96 0.89 0.86 0.75 0.83 0.78 0.77 0.66 0.76 0.73 0.71 0.65 0.71 0.70 0.69 0.60 35.52 40.31 45.24 54.20 25.87 28.45 30.65 36.34 16.04 16.91 19.53 23.78 18.81 25.27 28.70 36.38 0.32 0.25 0.20 0.18 Pressure) 0.49 0.39 0.35 0.32 200 bar (ORG Injection 0.92 0.58 0.79 0.52 0.76 0.49 0.69 0.46 1.80 1.97 2.21 2.80 1.55 1.70 1.91 2.32 1.23 1.31 1.45 1.83 1.19 1.28 1.39 1.82 0.39 0.32 0.26 0.21 0.40 0.33 0.31 0.26

0.96 0.88 0.84 0.74 0.76 0.70 0.68 0.59 0.70 0.66 0.65 0.55 0.67 0.65 0.64 0.54 45.70 51.34 60.92 73.61 30.37 33.51 36.23 43.66 18.74 20.21 23.01 28.08 21.83 27.99 31.53 39.48 0.48 0.37 0.32 0.30 0.74 0.63 0.56 0.48 M0 M5 M10 M15

0.42 0.38 0.31 0.26

Figure 2. The changes in the emissions relative to M0 at 5 N m load.

5Nm

0.38 0.34 0.26 0.21

0.36 0.29 0.23 0.21

0.37 0.30 0.24 0.19

1.23 1.32 1.44 1.87

1.28 1.37 1.49 1.92

1.70 1.90 2.14 2.62

2.08 2.37 2.63 3.32

0.81 0.68 0.64 0.58

220 bar 0.56 0.49 0.45 0.40

0.30 0.22 0.18 0.15

15 N m 15 N m 10 N m 20 N m 15 N m

CO

10 N m 5Nm 20 N m 15 N m 10 N m

Canakci et al.

fuel type

10 N m

15 N m

20 N m

5Nm

NOx UHC

emissions (g/kWh) except SN

5Nm

CO2

20 N m

5Nm

10 N m

SN

20 N m


Table 5. The Emission Results of the Fuels

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the fuel air mixing in the combustion chamber was more excellent, so the UHC emissions was obtained less than that of the low-injection pressure.25,42 Nitrogen Oxides (NOx) Emissions. The most troublesome emissions from CI engines are NOx emissions. The oxides of nitrogen in the exhaust emissions contain nitric oxide (NO) and nitrogen dioxides (NO2). The formation of NOx is highly dependent on in-cylinder temperatures, the oxygen concentration and residence time for the reaction to take place.43 The results showed that increasing methanol ratio in the fuel blend raised NOx emissions. For instance, as presented in Figure 2b, the change in NOx emissions was compared with M0, and it showed that NOx boosted by 7.6, 16.8, and 53% for M5, M10, and M15, respectively, at 5 N m load and ORG injection pressure. The oxygen content of methanol may provide some advantageous (postflame oxidation, flame speed etc.) during the air-fuel interactions, particularly in the fuel-rich region. Indeed, it was evident proof of the oxygen content of methanol enhanced the hydrocarbon oxidation since the measured CO, UHC emissions, and smoke opacity with use of the methanol significantly decreased. However, the oxygen content of methanol is an important factor in the NOx formation, because it causes higher local temperatures due to excess hydrocarbon oxidation. On the other hand, the LHV of methanol is nearly 2 times lower than diesel fuel and latent heat of vaporization of methanol is about 4 times greater than diesel fuel, which decreases peak (42) Sayin, C.; Hosoz, M.; Canakci, M.; Kilicaslan, I. Int. J. Energy Res. 2007, 31, 259–273.



temperature in the cylinder. However, as shown in Figure 6, the exhaust temperature increased with increasing methanol ratio in the fuel mixture. It is clear from the figure that cetane number and oxygen content are more effective than LHV and latent heat of vaporization with regard to increasing peak temperature in the cylinder. Therefore, the concentration of NOx increased as the methanol content was increased in the fuel blend.44 Different from CO and UHC emissions, NOx emissions increased with increasing engine load. For instance, comparing Figure 3a and Figure 5a, it was discovered that NOx emissions increased by 10% for M10 when the engine load was increased from 10 to 20 N m constant loads at the increased injection pressure. Figure 3a-c illustrate the percent change in NOx emissions with different methanol blends at different pressures for 10 N m load. As demonstrated in the figures for M5, decreased injection pressure reduced the NOx emission by 1.2%, and increased injection pressure increased the NOx emission by 4.5% when compared with ORG injection pressure. When the injection pressure was decreased, it was observed that NOx emissions (43) Agarwal, A. K. Prog. Energy Combust. Sci. 2007, 33, 233–271. (44) Nwafor, O. M. I.; Rice, G.; Ogbonna, A. I. Renewable Energy 2000, 21, 433–44. (45) Pukrakek, W. W. Engineering Fundamentals of the Internal Combustion Engine; Simon and Schuster Company: USA, 1997; p 278. (46) Canakci, M. Idealized Engine Emissions Resulting from the Combustion of Isooctane Supplemented with Hydrogen; MSc Thesis, Vanderbilt University: TN, 1996; pp 25-26. (47) Heywood, J. B. Internal Combustion Engines; Mc-Graw Hill: USA, 1984; pp 592-593.



diminished for the all fuel mixtures. Increasing the injection pressures decreased the particle diameter and caused the diesel-methanol fuel spray to vaporize quickly. However, the liquid fuel cannot penetrate deeply into the combustion chamber. So, higher injection pressure initially generates faster combustion rates, resulting in higher temperatures. As a consequence, NOx concentrations start to increase. The obtained exhaust gas temperatures, shown in Figure 6, confirm this statement.45 Smoke Number (SN). The emitted particulate matter is essentially composed of soot, though some hydrocarbons, generally referred to as a soluble organic fraction (SOF) of the particulate emissions, are also adsorbed on the particle surface or simply emitted as liquid droplets. Smoke opacity formation occurs at the extreme air deficiency. This air or oxygen deficiency is present locally inside diesel engines. It increases as the air-fuel ratio decreases. Soot is produced by oxygendeficient thermal cracking of long-chain molecules. The change in SNs compared to M0 implied that they diminished by 8, 10.5, and 23% for M5, M10, and M15, respectively, at 15 N m load and increased injection pressure as illustrated in Figure 4a. The presence of atomic bound oxygen in methanol satisfies positive chemical control over soot formation. The tendency to generate soot by the fuel-dense region inside a diesel diffusion flame sheath is reduced, so that soot-free spray combustion could be achieved.27 The formation of smoke is most strongly dependent on the engine load. As the load increases, more fuel is injected, and this increases the formation of smoke.46 The results acquired in this study supported this statement. For example, in com-

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Figure 6. Exhaust gas temperatures at different engine loads and ORG injection pressure.

parison with Figures 2c and 5c, it was shown that the change of SNs increased by 7% for M10 as the engine load increased from 5 to 20 N m at decreased injection pressure. Increasing the injection pressure decreases the smoke number. When injection pressure is boosted, fuel particle diameters will become smaller. Because formation of mixing fuel to air becomes better through injection period, smoke opacity will be less.29 Figure 5a-c presents the percent change in SNs at different injection pressures for a 20 N m load. As seen in these figures, for M15,

Canakci et al.

increased injection pressure lowered the SN 2.1%, and decreased injection pressure increased it 1.3% compared to ORG injection pressure for M15. Carbon Monoxide (CO) Emissions. CO is a colorless, odorless, poisonous gas, and it must be restricted. CO results from incomplete combustion of fuel and is emitted directly from vehicle tailpipes. Besides the ideal combustion process that combines carbon (C) and oxygen (O2) to CO2, incomplete combustion of carbon leads to the formation of CO. The formation of CO takes place when the oxygen presents during combustion is insufficient to form CO2.47 Regarding the influence of different fuel blends on CO emissions, it was found that increasing the methanol ratio in the blend reduced CO emissions. In comparison with M0, the change in CO emissions were 14.86, 24.32, and 35.13% for M5, M10, and M15, respectively, at 5 N m load and increased injection pressure, as illustrated in Figure 2a. Methanol is an oxygenated fuel and brings about more complete combustion; thus, CO emissions reduce in the exhaust.48 Increased engine load reduced CO emission gradually. When the engine load increases, combustion temperature increases and CO emissions begin to reduce. For example, compared with Figures 2a and 5a, it was shown that the percent change in CO emissions decreased by 10.61% for M5 as the engine load increased from 5 to 20 N m constant load at increased injection pressure. Figure 4a-c demonstrates the percent change in CO emissions with different methanol blends at different injection pressures for 20 N m load. From these figures, it was concluded that increased injection pressure reduced the CO emission by 2.51% and decreased injection pressure boosted the CO emission by 1.94% compared to ORG injection pressure for M15. The increasing injection pressure caused the good fuel-air mixing and easy and complete combustion of the smaller droplets. These effects lead to reduced CO emissions.49 Carbon Dioxide (CO2) Emissions. CO2 emissions are released into the atmosphere when fuel is completely burned in an engine. As demonstrated in the Figure 3a, when the methanol amount increased in the fuel mixture, the CO and UHC emissions decreased. The percent change in CO2 had an opposite behavior when compared with the CO concentrations, and this was due to improving the combustion process as a result of oxygen content in methanol. The maximum increase in the change of CO2 was obtained at 12.3, 33.3, and 61%, for M5, M10, and M15, respectively, compared to M0 at 10 N m engine load and increased injection pressure. In this study, CO2 emissions increased with the increased injection pressure for the all fuel mixtures. As shown in Figure 5a-c for M15, increased injection pressure augmented the change in CO2 by 7.8%, and decreased injection pressure diminished it by 6% compared to ORG injection pressure. Combustion Analysis. Figures 7 and 8 illustrate the cylinder gas pressures, and Figures 10 and 11 display the heat release rates for different fuels and ORG injection pressure at 20 and 10 N m loads, respectively. Figures 9 and 12 show the cylinder gas pressures and heat release rates for M0 and M15 at different injection pressure and 10 N m load. The measured start of fuel combustion and calculated ignition delay for each fuel are shown in Table 6 for different injection pressure. Peak Cylinder Gas Pressure. To analyze the cylinder gas pressure, the pressure data of 100 cycles with a resolution of 0.75 °CA were averaged and then used. The cylinder gas (48) Abdel-Rahman, A. A. Int. J. Energy Res. 1998, 22, 483–513. (49) Ajav, E. A.; Singh, B.; Bhattacharya, T. K. Biomass Bioenergy 1998, 15 (6), 493–502.



Figure 7. Cylinder gas pressure versus CA at 20 N m load and ORG injection pressure.

Figure 8. Cylinder gas pressure versus CA at 10 N m load and ORG injection pressure.

Figure 9. Cylinder gas pressure versus CA at different injection pressure and 10 N m load.

pressure variation is given in Figure 8 with respect to crank angle at 10 N m load and ORG injection pressure. As shown in the figure, the peak cylinder gas pressure shows an increase with the increase of methanol fraction in the fuel blend. The peak cylinder pressure occurred at 7.45 (at 12.48 °CA ATDC), 7.63 (at 12.34 °CA ATDC), 7.64 (at 11.04 °CA ATDC), and 7.65 (at 10.75 °CA ATDC) MPa for M0, M5, M10, and M15

at 10 N m load and ORG injection pressure, respectively. Since peak gas cylinder pressure is related to the amount of prepared fuel within the ignition delay period (premixed burning phase) and the lowering of gas temperature due to fuel evaporation, an increase in the methanol fraction will increase the fraction of fuel in the premixed combustion phase and then causes an increase in the peak cylinder pressure.50

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Canakci et al.

Figure 10. Heat release rate versus CA at 20 N m load and ORG injection pressure.

Figure 11. Heat release rate versus CA at 10 N m load and ORG injection pressure.

Figure 12. Heat release rate versus CA at different injection pressure and 10 N m load.

Comparing Figures 7 and 8, it is stated that the cylinder gas pressure decreased with increasing engine load. Experimental results indicate that the decrease in the cylinder pressure was approximately 5.6% for M5 when the engine load was increased from 10 to 20 N m. The peak pressure rise corresponds to the large amount of fuel burnt in the premixed combustion phase

and also earlier start of combustion at low load compared to high load. These behaviors provided an increase in the maximum cylinder gas pressure at low loads. The locations of maximum cylinder gas pressure closed to TDC at 10 N m engine load since starting the fuel injection taken place earlier than that of 20 N m.



Table 6. Combustion Characteristics of the Fuels at Different Injection Pressure M10

M15

200 bar (ORG injection pressure), 20 N peak cylinder gas pressure (MPa) 7.452 7.231 start of combustion (° BTDC) 12.05 11.04 ignition delay (°) 7.95 8.16

M0

M5

m 7.277 11.33 8.67

6.765 13.34 8.95

180 bar, 20 N m peak cylinder gas pressure (MPa) 7.712 7.559 start of combustion (° BTDC) 9.89 11.47 ignition delay (°) 10.11 8.53

7.402 12.48 7.52

7.206 12.49 7.55


7.634 13.63 6.37

7.559 12.91 7.09

200 bar (ORG injection pressure), 10 N peak cylinder gas pressure (MPa) 7.448 7.662 start of combustion (° BTDC) 12.48 12.34 ignition delay (°) 7.52 7.66

m 7.633 11.04 8.96

7.637 10.75 9.25


7.589 12.05 7.95

7.477 11.47 8.53


7.335 13.06 6.94

7.714 12.48 7.52

the ignition delay period and leads to increasing in the premixed heat release.27,51 Conclusions

Figure 9 displays a comparison of the changes in the cylinder gas pressures with respect to crank angle obtained for M0 and M15 at different injection pressures and 10 N m load. As demonstrated in this figure, the peak cylinder gas pressure was obtained 7.72 (at 6.36 °CA ATDC), 7.64 (at 7.08 °CA ATDC), and 7.45 (at 12.48 °CA ATDC), 7.63 (at 6.65 °CA ATDC) MPa at 10 N m loads for increased, ORG, and decreased injection pressures, respectively. The trend was such that, as injection pressure is increased, the atomization at the nozzle outlet is found to be enhanced. This results in a more-distributed vapor phase and better combustion.26 Heat Release Rate (HRR). As shown in the Figure 10, HRR generally decreases with the increase of methanol amount in the fuel blend. The maximum obtained HRR was 29.57 (at 22.26 °CA BTDC), 29.62 (at 23.86 °CA BTDC), 28.46 (at 21.12 °CA BTDC), and 27.46 (at 23.02 °CA BTDC) kJ/deg for M0, M5, M10, and M15 at 20 N m load and ORG injection pressure, respectively. Because methanol does not evaporate as easily as diesel fuel, the ignition delay increases with increasing methanol ratio, as seen in Table 6. The increase in the ignition delay may cause more fuel to be burned in the premixed burned phase and increase the heat release rate. Conversely, the LHV of methanol is lower than diesel fuel, which reduces HRR. It is clear from the figure that the LHV of methanol is more efficient than heat of vaporization with regard to HRR.26 Figures 10 and 11 show the HRR for different fuel blends and ORG injection pressure at 20 and 10 N m load, respectively. As seen in the figures, HRR increased with the rise of engine load because of the increase in the quantity of fuel injected. Figure 12 demonstrates a comparison of the changes in the cylinder gas pressures with respect to crank angle obtained for M0 and M15 at different injection pressures and 10 N m load. The ignition delay in a diesel engine is defined as the time between the start of fuel injection and the start of combustion. For M0 and M15, the ignition delay was decreased with the increasing injection pressure, as seen in Table 6. With increasing injection pressure, the ignition delay decreases because of the effects of active combustion caused by smaller particles of atomized fuel. This leads to greater accumulation of the fuel in

In this research, the effect of injection pressure on the combustion characteristics, engine performance, and exhaust emissions of a diesel engine has been experimentally investigated when using methanol-blended diesel fuel. The following conclusions can be drawn from the present paper: (1) The results stated that smoke number, CO, and UHC emissions decreased with the methanol amount in the fuel blend, and NOx and CO2 emissions increased because of the improved combustion. By using methanol-blended diesel fuels, the smoke number, CO, and UHC emissions were reduced 7-32, 10-50, and 11-46%, while CO2 and NOx emissions increased 5-51 and 7-35%, respectively, depending on the engine test conditions. Increasing the amount of methanol in the fuel blend caused higher peak temperature in the cylinder. This effect increased NOx emissions. (2) The results proved that increasing the ratio of methanol in the fuel blend leads to an increase in BSFC and BSEC and to a decrease in BTE. This is may be the result of the LHV of methanol, which is noticeably lower than that of the diesel fuel. The peak cylinder pressure and HRR reduced with increasing methanol ratio in the fuel blend. The lower cetane number with methanol supplement caused some increase in the ignition delay. Increasing ignition delay lead to deteriorating combustion, and the peak cylinder pressure decreased. The increase in the ignition delay augmented the rate of pressure rise and decreased the peak cylinder gas pressure. (3) With respect to injection pressure, the test results showed that, with increasing the injection pressure, NOx and CO2 emissions increased while smoke number, CO, and UHC emissions decreased. This lead to an increase in the CO2 emissions. Increasing the injection pressure caused an earlier start of combustion relative to the TDC. As a result of this, the cylinder charge, being compressed as the piston moves to the TDC, had relatively higher temperatures and so lowered the UHC emissions and increased NOx emissions. With the increasing injection pressure, the best results were attained for the UHC and CO emissions at 20 N m, and the best results were obtained for the smoke number at 5 N m load. At these situations, smoke number, CO, and UHC emissions were found to be 54%, 0.15 g/kWh, and 0.19 g/kWh, for M15, respectively. On the other hand, decreasing the injection pressure to 5 N m load offered the minimum results for NOx and CO2 emissions. At these working conditions, NOx and CO2 emissions were found to be 1.23 g/kWh and 21.83 g/kWh, respectively. (4) The best results for BSFC, BSEC, and BTE were obtained at the ORG injection pressure compared to the other injection pressures. When the injection pressure is low, the ignition delay period will increase, and this causes a reduction in the engine output power. Consequently, fuel consumption per output power will increase. But, if the injection pressure is high, the ignition delay period will become lower. So, the possibilities of homogeneous mixing are reduced and combustion efficiency decreases. In addition, increased injection pressure increased (50) Canakci, M.; Sayin, C.; Gumus, M. Energy Fuels 2008, 22, 3709– 3723. (51) Ozsezen, A. N.; Canakci, M.; Sayin, C. Energy Fuels 2008, 22, 1297–1305. (52) Lombardini. Engine Technical Specification; Turkey, 2000 [In Turkish]. (53) Merck. Product specification; Germany: 2006. (54) Tupras. Product Specification; Turkey: 2005 (In Turkish).

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peak cylinder pressure and HRR because of the decrease in the ignition delay. Acknowledgment. The authors would like to acknowledge Marmara University for providing financial support during the project (Project BSE-075/131102) and Kocaeli University Alternative Fuels R&D Center for their contribution in the experiments.

Nomenclature BTE ) brake thermal efficiency, (%) BSEC ) brake-specific energy consumption, (MJ/kWh) BSFC ) brake-specific fuel consumption, (g/kWh)

Canakci et al. CA ) crank angle CI ) compression ignition CO ) carbon monoxide CO2 ) carbon dioxide UHC ) unburned hydrocarbon LHV ) lower heating value, (kJ/kg) NOx ) nitrogen oxides ORG ) original ppm ) parts per million rpm ) revolution per minute TDC ) top dead center EF900060S

Effect of Injection Pressure on the Combustion ... - ACS Publications

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