Exhaust Emissions and Combustion Characteristics of a Direct Injection

Energy & Fuels 2008, 22, 3709–3723

3709

Exhaust Emissions and Combustion Characteristics of a Direct Injection (DI) Diesel Engine Fueled with Methanol-Diesel Fuel Blends at Different Injection Timings Mustafa Canakci,*,†,‡ Cenk Sayin,§ and Metin Gumus§ Department of Mechanical Education, Kocaeli UniVersity, Izmit 41380, Turkey, AlternatiVe Fuels Research and DeVelopment Center, Kocaeli UniVersity, Izmit 41040, Turkey, and Department of Mechanical Education, Marmara UniVersity, Istanbul 34722, Turkey ReceiVed May 26, 2008. ReVised Manuscript ReceiVed August 20, 2008

In the recent years, environmental concerns and depletion in petroleum resources have forced researchers to concentrate on finding renewable alternatives to conventional petroleum fuels. Therefore, alcohols as renewable and alternative energy sources for the diesel engines gain importance. For this reason, in this study, the performance, exhaust emissions, and combustion characteristics of a single cylinder diesel engine have been experimentally investigated under different injection timings when methanol-blended diesel fuel was used from 0 to 15%, with an increment of 5%. The tests were conducted at three different injection timings (15°, 20°, and 25 °CA BTDC) by changing the thickness of advance shim. All tests were conducted at four different loads (5, 10, 15, and 20 Nm) at constant engine speed of 2200 rpm. The experimental test results showed that BSFC, BSEC, combustion efficiency, and NOx and CO2 emissions increased as BTE, rate of heat release, peak cylinder pressure, smoke number, and CO and UHC emissions decreased with an increasing amount of methanol in the fuel blend. In comparison to the values at the original injection timing (20 °CA BTDC), the values at the retarded injection timing (15 °CA BTDC) of peak cylinder pressure, rate of heat release, combustion efficiency, and NOx and CO2 emissions decreased, while smoke number and UHC and CO emissions increased at all test conditions. On the other hand, The advanced injection timing (25 °CA BTDC), smoke number, and UHC and CO emissions diminished and peak cylinder pressure, rate of heat release, combustion efficiency, and NOx and CO2 emissions increased at all test conditions. In terms of BSFC, BSEC, and BTE, retarded and advanced injection timings gave negative results in all fuel blends compared to original injection timing.

Introduction Because of their fuel economy and high reliability, compression-ignition (CI) engines known as diesel engines have been penetrating a number of markets around the world. The existing CI engines operate with conventional diesel fuel derived from crude oil. It is well-known that the world petroleum resources are limited and the production of crude oil is becoming more difficult and expensive. On the other hand, because of the growing concern over possible adverse health effects caused by diesel emissions, the pollutions including unburned hydrocarbons (UHC), carbon monoxide (CO), nitrogen oxides (NOx), and particulate matter [quantified as smoke number (SN)] have been regulated by laws in many developed countries. In the last few years, many studies on the IC engines have been carried out with the aim of reducing exhaust emissions by changing operating parameters, such as valve timing, injection timing, and atomization rate. At the same time, depletion of fossil fuels and environmental considerations have led to investigations on renewable fuels, such as methanol, ethanol, hydrogen, and biodiesel.1-5 * To whom correspondence should be addressed. Telephone: +90-2623032285. Fax: +90-262-3032203. E-mail: [email protected]. † Department of Mechanical Education, Kocaeli University. ‡ Alternative Fuels Research and Development Center, Kocaeli University. § Marmara University. (1) Durgun, O.; Ayvaz, Y. The use of diesel fuel-gasoline blends in diesel engines. Proceedings of the 1st International Energy and Environment Symposium, Turkey, 1996; pp 9105-9120.

Alcohols (ethanol and methanol) have been considered as alternative fuels for diesel engines. Methanol is manufactured from any material that can be decomposed into CO (or CO2) and hydrogen. In this regard, it may be produced from the sources, which are independent from petroleum. The primary feedstocks for methanol production are natural gas, lignite coal, and renewable sources, such as wood, agricultural materials biomass, waste biomass, and municipal wastes.6-8 Diesel fuel is a complex mixture of a large number of hydrocarbons (such as C3-C25 hydrocarbons). For this reason, its fuel properties can change depending upon the proportion of hydrocarbon types used in the fuel mixture. However, methanol (CH3OH) is a simple compound. It contains an oxygen atom, so that it can be viewed as a partially oxidized hydrocarbon fuel. It has a lower heating value than diesel fuel; therefore, much more fuel is needed to obtain the same performance as that of a diesel-fueled engine. Its high stoichiometric fuel/air ratio, high oxygen content, and high H/C ratio may be beneficial for improving the combustion and reducing (2) Yuksel, F.; Yuksel, B. Renewable Energy 2004, 29, 1181–1191. (3) Asfar, K. R.; Hamed, H. Energy ConVers. Manage. 1998, 39 (10), 1081–1093. (4) Chao, M. R.; Lin, C. T.; Chao, H. R.; Chang, F. H.; Chen, C. B. Sci. Total EnViron. 2001, 279, 167–179. (5) Cinar, C.; Topgul, T.; Ciniviz, M.; Hasimoglu, C. Appl. Therm. Eng. 2005, 25, 1854–1862. (6) Thring, R. H. SAE Tech. Pap. 831685, 1983. (7) Chmielniak, T.; Sciazko, M. Appl. Energy 2003, 74, 393–403. (8) Ristinen, R.; Kraushaar, J. Energy and EnVironment; John Wiley and Sons: New York, 2006; pp 268-269.

10.1021/ef800398r CCC: $40.75  2008 American Chemical Society Published on Web 09/27/2008

3710 Energy & Fuels, Vol. 22, No. 6, 2008

the soot and smoke in diesel engines. Methanol has a higher latent heat of vaporization than diesel fuel. It extracts much more heat as it vaporizes, and it can lead to a cooling effect on the cylinder charge. Therefore, the cylinder gas temperature can be decreased because of its cooling effect on the charge, and emissions of NOx may be reduced. On the other hand, methanol has poor ignition behavior because of its low cetane number, high latent heat of vaporization, and high ignition temperature. Thus, it can cause a longer ignition delay. The autoignition temperature of methanol is higher than that of diesel fuel, which makes it safer for transportation and storage. However, it has a lower flash point than that of diesel fuel, which is a disadvantage for safety.9-11 The possible benefits and shortcomings of methanol as a fuel for CI engines are summarized above. Problems concerning the use of methanol in diesel engines can be different, which are briefly described below. The simplest method of using methanol in a CI engine is to blend it with diesel. The most important problem is the phase separation. Using a mixer inside the fuel tank can prevent this problem. Moreover, an ignition improver, such as diethyl ether, can be added to the blended fuel to compensate for the cetane number.12 The exhaust emissions from diesel vehicles are very complex mixtures containing many types of pollutants, which can be found in different forms, such as particulate, semi-volatile, and gaseous phases. Using methanol-blended diesel fuel can reduce the air pollution. Therefore, many researchers have studied the influence of this alternative fuel on the engine performance and exhaust emissions of IC engines. Huang et al.,13 for instance, used various blend rates of methanol-diesel fuels in the engine tests. The results indicated that the increase of methanol content decreased smoke number and CO and UHC emissions but increased brake-specific fuel consumption (BSFC) and NOx emissions. Lin and Chao14 investigated the effect of the methanolcontaining additive (MCA) on the biological characteristics of diesel exhaust emissions. The engine was tested on a series of diesel fuels blended with five additive levels (0, 5, 8, 10, and 15% of MCA by volume). The result showed that the MCA additive moderately lowered the toxicity levels of particleassociated (SOF) samples but generally increased the vaporphase (XOC) associated toxicity. Bayraktar15 studied the effect of methanol-blended diesel fuel between 2.5 and 15 vol % on the engine performance and found that the methanol could reduce the effective power and brake thermal efficiency (BTE) to some degree and moderately increase BSFC. Cheng et al.16 investigated the performance and exhaust emissions of a four-cylinder diesel engine operating on biodiesel with methanol in either the blended or fumigation mode. They compared the results to those operating on pure biodiesel and pure diesel fuel. Experiments were performed on a four-cylinder naturally aspirated direct injection (DI) diesel engine operating at a constant speed of 1800 rpm with five different engine loads. (9) Kowalecwicz, A. Proc. Inst. Mech. Eng. 1993, 207, 43–52. (10) Wagner, T. O.; Gray, D. S.; Zarah, B. Y.; Kozinski, A. A. SAE Tech. Pap. 790429, 1979. (11) Adelman, H. SAE Tech. Pap. 790956, 1979. (12) Murayama, T.; Miyamoto, N.; Yamada, T.; Kawashima, J. I. SAE Tech. Pap. 821113, 1982. (13) Huang, Z. H.; Lu, H. B.; Jiang, D. M.; Zhang, J. Q.; Wang, X. B. Proc. Inst. Mech. Eng., Part D 2004, 218, 435–447. (14) Lin, T. C.; Chao, M. R. Sci. Total EnViron. 2002, 284, 61–74. (15) Bayraktar, H. Fuel 2008, 87, 158–164. (16) Cheng, C. H.; Cheung, C. S.; Chan, T. L.; Lee, S. C.; Yao, C. D.; Tsang, K. S. Fuel 2008, 87, 1870–1879.

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The results indicated reductions in CO2, NOx, and particulate mass emissions and reductions in the mean particle diameter, in both cases, compared to diesel fuel. Chao et al.4 investigated the effect of the MCA on the regulated and unregulated emissions from a diesel engine. The engine was tested on a series of diesel fuels blended with five additive levels (0, 5, 8, 10, and 15% of MCA by volume). Results showed that MCA addition slightly decreased particulate matter (PM) emissions but generally increased both UHC and CO emissions. Cheung et al.17 studied the effects of fumigation methanol on the emissions of a diesel engine fueled with biodiesel as the baseline fuel. The biodiesel used in this study was converted from waste cooking oil. Experiments were performed on a fourcylinder naturally aspirated diesel engine operating at a constant speed of 1800 rpm for three engine loads. The results indicated no significant change in BTE and CO2 emission, an increase in both CO and UHC emissions, and a decrease in both NOx and PM emissions. The Lubrizol Corporation, in conjunction with the Lovelace Respiratory Research Institute and several subcontracting laboratories, recently conducted a health assessment of the combustion emissions of Puri NOx diesel fuel emulsion (diesel-water-methanol) in rodents. Combustion emissions from either of two, 2002 model Cummins 5.9 L ISB engines were diluted with charcoal-filtered air to expose concentrations of 125, 250, and 500 mg of total PM/m3. The engines were operated on a continuous, repeating, heavy-duty certification cycle (U.S. Code of Federal Regulations, Title 40, Chapter I). NO and PM were reduced when engines were operated on Puri NOx versus California Air Resources Board diesel fuel under these conditions.18 In a different study,19 a stabilized methanol-diesel blend was developed and combustion characteristics and heat release analysis were carried out in a CI engine. According to the experimental results, increasing the methanol mass fraction in the methanol-diesel fuel blends resulted in an increase in the heat release rate at the premixed burning phase and shortened the combustion duration at the diffusive burning phase. Yao et al.20 investigated the controlling strategies of homogeneous charge compression ignition (HCCI) fueled by dimethyl ether (DME) and methanol. The experimental work was carried out on a modified single-cylinder diesel engine, which was fitted with port injection of DME and methanol dual fuel. The results showed that the exhaust gas recirculation (EGR) rate and DME percentage are two important parameters to control the HCCI combustion process. For a diesel engine, fuel injection timing is a major parameter that affects the combustion and exhaust emissions. The state of air into which the fuel injected changes as the injection timing is varied and, thus, ignition delay will vary. If the injection starts earlier, the initial air temperature and pressure are lower; therefore, the ignition delay will increase. If the injection starts later (when the piston is closer to TDC), the temperature and pressure are initially slightly higher and a decrease in ignition delay results. Hence, variation in injection timing has a strong (17) Cheung, C. S.; Cheng, C.; Chan, T. L.; Lee, S. C.; Yao, C.; Tsang, K. S. Energy Fuels 2008, 22, 906–914. (18) Reed, M. D.; Blair, L. F.; Daly, I.; Gigliotti, A. P.; Gudi, R.; Mercieca, M. D.; McDonald, J. D.; O’Callaghan, J. P.; Seilkop, S. K.; Ronsko, N. L.; Wagner, V. O.; Kraska, R. C. Toxicol. Ind. Health 2006, 22, 65–85. (19) Huang, Z. H.; Lu, H. B.; Jiang, D. M.; Zhang, J. Q.; Wang, X. B. J. Automot. Eng. 2004, 218, 1011–1024. (20) Yao, M.; Chen, Z.; Zheng, Z.; Zhang, B.; Xing, Y. Fuel 2006, 85, 2046–2056.

DI Diesel Engine with Methanol-Diesel Fuel Blends

effect on the engine performance and exhaust emissions, especially on the BSFC, BTE, and NOx emissions, because of changing the maximum pressure and temperature in the engine cylinder.21,22 Several studies have shown that the injection timing affects the engine performance and exhaust emissions of CI engines. Parlak et al.23 studied the influence of injection timing on the NOx emissions and BSFC of a low-heat rejection (LHR) IDI diesel engine using diesel fuel. The tests were conducted with variable loads at the engine speeds of 1000, 1400, 1800, and 2000 rpm and the static injection timing of 38°, 36°, 34°, and 32° crank angle (CA) before top dead center (BTDC). After the load tests were conducted for the original diesel engine, the same test order was adopted for the LHR engine. When the LHR engine was operated with the injection timing of 38 °CA BTDC, which is the original start of injection timing of the engine, the NOx emission increased about 15%. However, when the injection timing was retarded to 34 °CA in the LHR case, some decreases in the exhaust emissions and BSFC were observed compared to those of the original case. Nwafor24 examined the effect of advanced injection timing on the engine performance and exhaust emissions of natural gas used as a primary fuel in a dual-fuel CI engine. The test engine had an original (ORG) start of injection timing of 30 °CA BTDC. The injection was advanced by 3.5° (i.e., 33.5 °CA BTDC). The results indicated that dual-fuel combustion produced higher HC emissions than that of pure diesel fuel. Significant reductions in CO and CO2 emissions were obtained when running the engine with the advanced injection timing. On the other hand, advanced injection timing caused a slight increase in BSFC and decrease in BTE. Sayin and Canakci25 investigated the influence of injection timing on the exhaust emissions and engine performance of a DI diesel engine using ethanol-blended diesel fuel from 0 to 15%, with an increment of 5%. The original start of injection timing of the engine was 27 °CA BTDC in that study. The tests were performed at five different injection timings (21°, 24°, 27°, 30°, and 33 °CA BTDC). The experimental test results showed that BSFC and emissions of NOx and CO2 increased as BTE and emissions of CO and HC decreased with an increasing amount of ethanol in the fuel mixture. When compared to the results of original injection timing (27 °CA BTDC), NOx and CO2 emissions increased and unburned HC and CO emissions decreased for the retarded injection timings (21° and 24 °CA BTDC) at the all test conditions. On the other hand, with the advanced injection timings (30° and 33 °CA BTDC), HC and CO emissions decreased and NOx and CO2 emissions increased. In terms of BSFC and BTE, retarded and advanced injection timings gave negative results for all engine speeds and loads as compared to the original injection timing in all fuel blends. From the literature review, the influence of injection timing on the exhaust emissions, combustion characteristics, and performance parameters of a diesel engine has not been clearly studied when using methanol-blended diesel fuel in a CI engine. Therefore, these topics need to be investigated to make up for the deficiency in the literature. For this reason, in the present study, the effects of both injection timing and methanol-blended (21) Heywood, J. B. Internal Combustion Engines; Mc-Graw Hill: New York, 1984; pp 546-547 and 882-383. (22) Borat, O.; Balci, M.; Surmen, A. Internal Combustion Engines; Gazi University Publishing: Turkey, 2000; pp 264-265 (in Turkish). (23) Parlak, A.; Yasar, H.; Hasimoglu, C.; Kolip, A. Appl. Therm. Eng. 2005, 25, 3042–3052. (24) Nwafor, O. M. I. Sadhana 2002, 27 (3), 375–382. (25) Sayin, C.; Canakci, M. Energy ConVers. Manage., in press, doi: 10.1016/j.enconman.2008.06.007.

Energy & Fuels, Vol. 22, No. 6, 2008 3711 Table 1. Technical Specifications of the Test Engine51 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 18 Nm at 2200 rpm 8 kW at 2000 rpm

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

accuracy

load speed time temperatures

(2 Nm (25 rpm (0.5% (1 °C

calculated results

uncertainty

power BSFC BSEC BTE

(2.55% (2.60% (2.60% (2.60%

diesel fuel on the engine performance, exhaust emissions, and combustion characteristics were experimentally investigated on a single-cylinder CI engine. Experimental Apparatus and Procedure The experiments were conducted on a single-cylinder, fourstroke, DI, and naturally aspirated CI engine. Details of the engine specification are shown in Table 1. The engine was loaded by an electrical dynamometer 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 by 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. The crankshaft position was obtained using a crankshaft angle sensor to determine the cylinder pressure as a function of the crank angle. The crank angle signal was obtained from an anglegenerating device mounted on the main shaft. The signal of the cylinder pressure was acquired for every 0.75 °CA, and the acquisition process covered 100 completed cycles. The engine oil temperature, coolant temperature, exhaust temperature, inlet air temperature were measured using K-type thermocouples. CO, CO2, and HC emissions were measured with an infrared gas analyzer (Bilsa Mod 210) with an accuracy of (0.001%, (0.01%, and (1 ppm, respectively. NOx emissions were recorded using an electrochemical gas analyzer (Kane-May Qintox KM9106) with an accuracy of (1 ppm. Smoke levels were obtained using a Bosch system with an accuracy of (0.1%. The analyzers were calibrated before the experiments. 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 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 (eurodiesel and methanol) were used. Euro-diesel was obtained from TUPRAS Petroleum Corporation. Methanol, with a purity of 99%, was purchased from a commercial supplier. The fuel properties are shown in Table 3. The euro-diesel was blended with methanol to obtain four different fuel blends from 0 to 15%, with an increment of 5%. The fuel blends were prepared just before starting the


Canakci et al.

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 (mm2/s, at 25 °C) stoichiometric air/fuel ratio stoichiometric fuel/air ratio heat of vaporization (MJ/kg)

methanol49

euro-diesel50

CH3OH 32 64.7 0.79 11 470 20.27 4 0.59 6.66 0.15 1.11

C14.34H24.75 196.8 287 0.83 78 235 42.74 56.5 3.35 14.28 0.07 0.27

ηcombustion ) LHVfuel - HNONO - HNO2NO2 - HCOCO - LHVfuelUHC LHVfuel

Calculation Methods Heat release analysis can yield valuable information about the effect of engine design changes, fuel injection system, fuel type, and engine operating conditions on the combustion process and engine performance.26 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:

1 γ (PdV) + (VdP) + Qw γ-1 γ-1

× 100 (2)

experiment to ensure that the fuel mixture was homogeneous. A mixer was also mounted inside the fuel tank to prevent phase separation. All data were collected after the engine stabilized. Each test was repeated 3 times. The values given in this study are the average of these three results. The experimental setup is shown in Figure 1. The original start of injection timing of the engine was 20 °CA BTDC. The thickness of one advance shim, located in connection place between the engine and fuel pump was 0.25 mm, and adding or removing one shim changed the injection timing by 5 °CA. Experiments were carried out in three different injection timings (15°, 20°, and 25 °CA BTDC) values. All tests were conducted at four different loads (5, 10, 15, and 20 Nm) 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.

Q)

instead, the LHV was used. Each enthalpy of formation calculation is mass-specific to the emission quantity measured during the engine testing.30 The enthalpy of formation for the measured exhaust emissions is used to calculate the combustion efficiency in accordance with the equation below

(1)

where Q is the apparent heat release rate (J), γ is the ratio of specific heats, which is calculated according to an empirical equation,27 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 on the basis of the Hohenberg correlation,28 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 upon temperature; leakage through the piston rings was neglected.29 Because the chemical energy of the fuel 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, 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; (26) Ghojel, J.; Honnery, D. Appl. Therm. Eng. 2005, 25, 2072–2085. (27) Brunt, M. F. J.; Rai, H.; Emtage, A. L. SAE Tech. Pap. 981052, 1998. (28) Hohenberg, G. H. SAE Tech. Pap. 790825, 1979. (29) Hayes, T. K.; Savage, L. D.; Sorenson, S. C. SAE Tech. Pap. 860029, 1986.

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

Results and Discussion Methanol-blended diesel fuel can decrease the pollutant emissions. However, to reach the emission reduction, it may require some modification on the engine. The injection timing has a significant effect on the performance and exhaust emissions of a CI engine. Therefore, in this study, the effects of injection timing and methanol-blended diesel fuel on the engine performance, exhaust emissions, and combustion characteristics were experimentally investigated on a single-cylinder CI engine. The experimental conditions were selected as follows: four engine loads (5, 10, 15, and 20 Nm), 2200 rpm constant speed, and three injection timings (15°, 20°, and 25 °CA BTDC). 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 in volume). Engine Performance. Table 4 shows the BSFC, BSEC, BTE, and combustion efficiency values and the percent changes in these parameters compared to M0 (diesel fuels), respectively, at different engine loads and injection timings. Brake-Specific Fuel Consumption (BSFC). The BSFC is defined as the ratio of the fuel consumption to the brake power. As shown in Table 4, the change in BSFC relative to M0 is 3.94, 7.67, and 18.26% for M5, M10, and M15, respectively, at 5 Nm and ORG injection timing. The results show that increasing the methanol ratio in the fuel mixture leads to an increase in BSFC. This behavior is attributed to the LHV of methanol, which is distinctly lower than that of the diesel fuel.31 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 2.07 times as the engine load increases from 5 to 20 Nm constant loads for M10. This decrease in BSFC could be explained by the fact that, as the engine load increased, the rate of increasing brake power was much more than that of the fuel consumption. As seen in the table, the minimum BSFC values were obtained at ORG injection timing for all fuel blends. When the injection timing was advanced and retarded compared to ORG injection timing, the change in BSFC was measured by 13.58 and 22.22%, respectively, for M10 at 10 Nm. With advancing (30) Canakci, M. Bioresour. Technol. 2007, 98, 1167–1175. (31) 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.


Energy & Fuels, Vol. 22, No. 6, 2008 3713

Figure 1. Experimental setup.

injection timing, the ignition delay will be longer and the speed of the flame will be shorter. These cause a reduction in engine output power. Therefore, fuel consumption per output power will increase. On the other hand, retarding injection timing means later combustion pressure rises only when the cylinder volume was expanding rapidly, and this reduces the effective pressure to do work.32 Brake-Specific Energy Consumption (BSEC). The BSEC is described as the product of BSFC and LHV. As shown in the Table 4, the BSEC increases with methanol content. Minimum BSEC was acquired at 5.05 MJ/kWh for M0, 5.08 MJ/kWh for M5, 5.10 MJ/kWh for M10, and 6.02 MJ/kWh for M15 at 20 Nm load and ORG injection timing, 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 2 times higher than that of methanol, as shown in Table 3. For these reasons, the effective power should decrease with the increase of the methanol amount in the fuel mixture. Thus, the engine needs to consume more heat to maintain the same amount of power output.33,34 The BSEC reduced as the engine load increased because of noticeably diminishing BSFC for all fuel blends and injection timing. The BSEC reduced by 14.70% as the engine load increased from 10 to 20 Nm constant loads for M15 at retarded injection timing. When the injection timing changed from ORG injection timing, BSEC values increased because of the increase in the energy requirement to sustain the same amount of power output at ORG injection timing. The increments for the advanced and the retarded injection timings were 9.92 and 13.29% for M0 at 10 Nm, respectively. Brake Thermal Efficiency (BTE). The BTE is defined as the ratio of the brake power to fuel consumption and LHV. As demonstrated in Table 4, in comparison to M0, the BTE decreased by 3.84, 14.25, and 34.17% for M5, M10, and M15, respectively, at 15 Nm and ORG injection timing. BTE indicates 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 (32) Sayin, C.; Kilicaslan, I.; Canakci, M.; Ozsezen, N. Appl. Therm. Eng. 2004, 8, 1315–1324. (33) Can, O.; Celikten, I.; Usta, N. Energy ConVers. Manage. 2004, 45, 2429–2440. (34) Rakopolous, C. D.; Kyristis, D. C. Energy 2001, 26, 705–722.

methanol amount increases in the fuel blend, the BSFC increases, because 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 timings.35 Increasing engine loads caused an increase in BTE values because of a noticeable decline in BSFC for all fuel blends and injection timings. The BTE increased by 16.46% as the engine load increased from 10 to 20 Nm at constant loads for M5 at advanced injection timing. The best results in terms of BTE were obtained at ORG injection timing. Retarded or advanced injection timing diminished BTE values by increasing BSFC. When the injection timing was advanced and retarded in comparison to ORG injection timing, BTE decreased by 19.44 and 31.02% for M10 at 20 Nm load, respectively. Combustion Efficiency. The change in combustion efficiency increased with an increasing methanol ratio in the fuel blend at all engine loads and injection timings, compared to M0. As shown in Table 4, the change in combustion efficiency increased by 0.15, 0.24, and 0.60% for M5, M10, and M15, respectively, at 20 Nm engine load and advanced injection timing 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 increased with an increasing engine load from 5 to 20 Nm for all test fuels because of better volumetric efficiency and atomization rate. The maximum combustion efficiency (99.88%) was obtained with advanced injection timing, followed by ORG (99.81%) and retarded injection timing (99.42%) for M15 at 20 Nm load. The increase in combustion efficiency with advancing the injection timing 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 CO, CO2, UHC, NOx, and smoke number (SN). Table 5 and Figures 2-5 show the emission values and percent changes in the emissions at different engine load and injection timing compared to M0, respectively. Carbon Monoxide (CO) Emissions. In general, while the engine is running under fuel-rich mixture conditions, the exhaust (35) Sayin, C.; Uslu, K. Int. J. Energy Res. 2008, 32, 1006–1015.

M15

M10

M5

M0

M15

M10

M5

M0

M15

M10

264.40 (-) 301.06 (13.8) 344.32 (30.22) 402.62 (53.78)

250.05 (-) 259.92 (3.94) 269.23 (7.67) 295.71 (18.26)

259.30 (-) 274.13 (5.71) 285.73 (10.19) 336.22 (29.66)

M0

M5

5 Nm

fuel type

148.71 (-) 177.19 (19.15) 198.75 (33.64) 262.60 (76.58)

131.27 (-) 139.65 (6.38) 162.16 (23.53) 194.16 (47.90)

144.23 (-) 160.44 (11.23) 184.66 (28.03) 213.78 (48.22)

10 Nm

144.15 (-) 161.98 (12.36) 186.54 (29.40) 215.14 (49.24)

127.45 (-) 132.35 (3.84) 151.48 (18.85) 172.70 (35.50)

147.12 (-) 154.90 (5.28) 169.24 (15.03) 202.44 (37.60)

15 Nm

BSFC (% change)

142.73 (-) 153.47 (7.52) 169.67 (18.87) 198.77 (39.26)

117.54 (-) 121.47 (3.34) 123.06 (4.69) 152.13 (29.42)

134.56 (-) 136.80 (1.66) 142.26 (5.72) 181.73 (35.05)

20 Nm

11.36 (-) 12.60 (10.91) 14.01 (23.32) 15.92 (40.14)

10.75 (-) 10.88 (1.20) 10.96 (1.95) 11.69 (8.74)

11.14 (-) 11.47 (2.96) 11.63 (4.39) 13.30 (19.38)

5 Nm

6.39 (-) 7.42 (16.11) 8.09 (26.60) 10.39 (62.59)

5.64 (-) 5.84 (3.54) 6.60 (17.02) 7.68 (36.17)

6.20 (-) 6.71 (8.22) 7.51 (21.12) 8.45 (36.29)

10 Nm

5 Nm

28.64 (-) 27.44 (-4.18) 23.93 (-16.44)) 22.83 (-20.28)

25.59 (-) 25.03 (-2.18) 22.21 (-13.20) 20.41 (-20.24)

15 °CA BTDC 6.12 9.12 (-) (-) 6.42 8.72 (4.90) (-4.38) 6.90 8.11 (11.27) (-11.07) 7.86 7.59 (28.43) (-16.77)

6.20 (-) 6.78 (9.35) 7.59 (22.41) 8.51 (37.25)

35.14 (-) 33.79 (-3.84) 30.13 (-14.25) 23.13 (-34.17)

15 Nm

20 °CA BTDC (ORG Injection Timing) 5.48 5.05 11.51 33.29 (-) (-) (-) (-) 5.54 5.08 11.13 32.58 (1.09) (0.59) (-3.30) (-2.13) 5.93 5.10 11.12 28.82 (8.21) (0.99) (-3.38) (-13.42) 6.83 6.02 10.81 21.78 (24.63) (19.20) (-6.08) (-34.57)

10 Nm 29.62 (-) 28.39 (-4.15) 27.33 (-7.73) 22.18 (-25.11)

25 °CA BTDC 5.78 9.91 (-) (-) 5.72 9.18 (-0.01) (-7.36) 5.80 8.71 (0.34) (-12.10) 7.19 8.22 (24.39) (-17.05)

20 Nm

30.31 (-) 27.92 (-7.88) 25.44 (-16.06) 25.12 (-17.12)

43.42 (-) 38.32 (-11.74) 36.88 (-15.06) 25.82 (-40.53)

32.88 (-) 30.42 (-7.48) 29.71 (-9.64) 24.11 (-26.67)

20 Nm

brake thermal efficiency (% change)

27.19 (-) 26.12 (-3.93) 25.78 (-5.18) 20.79 (-23.53)

6.32 (-) 6.48 (2.53) 6.89 (9.01) 8.01 (26.74)

15 Nm

BSEC (% change)

Table 4. BSFC (g/kWh), BSEC (MJ/kWh), and Thermal and Combustion Efficiencies of the Engine

98.16 (-) 98.27 (-0.11) 98.45 (-0.29) 98.90 (-0.75)

98.94 (-) 99.03 (0.09) 99.04 (0.10) 99.19 (0.25)

99.11 (-) 99.26 (0.15) 99.34 (0.23) 99.53 (0.42)

5 Nm

98.35 (-) 98.46 (-0.11) 98.74 (-0.39) 99.11 (-0.77)

99.02 (-) 99.14 (0.12) 99.15 (0.13) 99.31 (0.29)

99.12 (-) 99.32 (0.20) 99.35 (0.23) 99.68 (0.56)

10 Nm

98.33 (-) 98.48 (-0.15) 98.84 (-0.51) 99.29 (-0.97)

99.12 (-) 99.23 (0.11) 99.32 (0.20) 99.59 (0.47)

99.19 (-) 99.34 (0.15) 99.54 (0.35) 99.67 (0.48)

15 Nm

98.44 (-) 98.59 (0.15) 99.03 (0.59) 99.42 (0.99)

99.17 (-) 99.38 (0.21) 99.41 0.24 99.81 (0.64)

99.28 (-) 99.43 (0.15) 99.52 (0.24) 99.88 (0.60)

20 Nm

combustion efficiency (% change)

3714 Energy & Fuels, Vol. 22, No. 6, 2008 Canakci et al.

0.99 0.88 0.87 0.82 0.86 0.77 0.74 0.71 0.76 0.72 0.70 0.67 0.72 0.71 0.70 0.65 0.42 0.35 0.29 0.23 0.45 0.38 0.32 0.26 1.04 0.86 0.86 0.77 0.38 0.33 0.29 0.26 0.95 0.82 0.72 0.63 M0 M5 M10 M15

0.44 0.40 0.35 0.29

0.31 0.26 0.22 0.17

42.65 43.13 44.84 48.03

24.68 25.13 28.03 31.37

26.94 29.18 31.02 36.13

21.16 24.35 26.76 34.28

1.36 1.44 1.56 1.66

0.73 0.81 0.90 1.02

0.84 0.96 1.17 1.22 15 °CA BTDC 0.70 0.76 0.83 0.95

0.50 0.43 0.38 0.30

0.96 0.84 0.82 0.79 0.83 0.73 0.70 0.66 0.76 0.70 0.69 0.65 0.71 0.69 0.67 0.62 0.35 0.30 0.25 0.20 0.37 0.31 0.28 0.21 0.80 0.72 0.65 0.57 0.27 0.23 0.19 0.17 0.71 0.60 0.53 0.46 M0 M5 M10 M15

0.33 0.29 0.24 0.21

0.23 0.19 0.17 0.14

46.54 47.55 48.99 53.05

25.14 26.78 30.13 32.51

26.44 28.76 32.44 36.65

20 °CA BTDC (ORG Injection Timing) 23.72 1.43 0.78 0.81 1.06 27.70 1.51 0.87 0.92 1.23 30.83 1.67 0.94 1.09 1.54 39.15 1.91 1.09 1.21 1.69

0.40 0.34 0.30 0.23

0.94 0.80 0.77 0.70 0.76 0.65 0.63 0.57 0.70 0.64 0.63 0.58 0.65 0.64 0.63 0.56 0.33 0.26 0.20 0.16 0.34 0.27 0.22 0.16 0.36 0.30 0.24 0.18 0.71 0.62 0.53 0.45 1.51 1.92 2.29 2.55 1.30 1.49 1.80 2.02 25 °CA BTDC 1.14 1.23 1.45 1.75 1.75 1.82 2.15 2.54 29.34 36.44 38.91 50.94 34.76 39.22 44.65 48.87 30.70 34.78 38.31 41.05 52.10 54.49 55.82 60.43 0.26 0.19 0.16 0.15 0.59 0.48 0.40 0.36 M0 M5 M10 M15

0.28 0.21 0.18 0.16

0.23 0.16 0.14 0.11

15 Nm 15 Nm 20 Nm 15 Nm

NOx

10 Nm

emissions (g/kWh) except SN

5 Nm 20 Nm 15 Nm 15 Nm 5 Nm fuel type

10 Nm

20 Nm

5 Nm

10 Nm

CO2 CO

Table 5. Emission Results of the Fuels

5 Nm

10 Nm

HC

20 Nm

5 Nm

10 Nm

SN

20 Nm



will contain a large amount of CO emission, because there is not sufficient oxygen to convert all of the carbon atoms of the fuel into CO2. Thus, the most important parameters that affect CO emissions are an insufficient amount of air and an insufficient time in the cycle for complete combustion.36 Concerning the effect of different fuels on CO emissions, it was uncovered that increasing the methanol ratio in the fuel blend lessened CO emissions. In comparison to M0, the change in CO emissions was 18.64, 32.20, and 38.98% for M5, M10, and M15, respectively, at 5 Nm load and advanced injection timing, as demonstrated in Figure 2a. Methanol is an oxygenated fuel and leads to more complete combustion; hence, CO emissions reduce in the exhaust. CO emission decreased gradually when the engine load increased. When the engine load increases, the combustion temperature increases and CO emissions start to decrease.37 For example, in comparison to Figure 3b and 4b, it was seen that the percent change in CO emissions diminished by 20.68% for M5 as the engine load increased from 10 to 15 Nm constant load at ORG injection timing. Parts a-c of Figure 3 illustrate that the percent change in CO emissions with different methanol blends at different injection timings for 10 Nm load. From these figures, it was concluded that advanced injection timing decreased the CO emission by 8.44% and retarded injection timing increased the CO emission by 6.82% compared to ORG injection timing for M10, respectively. The advanced injection timing produced the higher cylinder temperature and increasing oxidation process between carbon and oxygen molecules. These lead to a decrease in the percent change in CO emissions.38 Unburned Hydrocarbon (UHC) Emissions. UHC emissions consist of fuel that is incompletely burned. The term HC means organic compounds in the gaseous state, solid HCs are part of the particulate matter. Typically, UHCs are a serious problem at low loads in CI engines. At low loads, the fuel is less apt to impinge on surfaces; however, 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.39 With regard to the effect of different methanol contents on UHC emission, it was found that increasing the methanol ratio in the fuel blend reduced UHC emissions. For instance, the UHC emissions compared to M0 at ORG injection timing decreased by 14, 24, and 40% for M5, M10, and M15, respectively, at 10 Nm load and retarded injection timing, as seen in Figure 3c. 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.40 UHC emissions lessened reasonably with increasing load, which was the same trend as with CO. For example, in comparison to Figure 3a to 5a, it was observed that the change in UHC emissions diminished by 6.66% for M10 as the engine load increased from 10 to 20 Nm constant loads at advanced injection timing. Parts a-c of Figure 5 show the change in UHC (36) Ganesan, V. Internal Combustion Engine; McGraw-Hill: New York, 1994; p 414. (37) Abdel-Rahman, A. A. Int. J. Energy Res. 1998, 22, 483–513. (38) Gumus, M. Renewable Energy, in press, doi: 10.1016/ j.renene.2008.02.005. (39) Canakci, M. Idealized engine emissions resulting from the combustion of isooctane supplemented with hydrogen. M.Sc. Thesis, Vanderbilt University, Nashville, TN, 1996; pp 25-26. (40) Alla, G. H.; Soliman, H. A.; Badr, O. H.; Rabbo, M. F. Energy ConVers. Manage. 2002, 43, 269–277.


Canakci et al.

Figure 2. Changes in the emissions relative to M0 at 5 Nm load.

emissions with different methanol blends at different injection timings for 20 Nm load compared to M0. From these figures, it was found that advanced injection timing caused a reduction in UHC emission by 8.66% and retarded injection timing boosted the UHC emission by 2.55% compared to ORG timing for M15, respectively. Advancing the injection timing caused an earlier start of combustion relative to the TDC. Because of this, the cylinder charge, being compressed as the piston moved to the TDC, had relatively higher temperatures and thus lowered the UHC emissions.41-43 Nitrogen Oxides (NOx) Emissions. In a diesel engine, the fuel distribution is non-uniform. The pollutant formation process is strongly dependent upon the changes in the fuel with time because of mixing. The oxides of nitrogen form in the high(41) Payri, F.; Bejanes, J.; Arregle, J.; Riesco, J. M. Oil Gas Sci. Technol. 2006, 2, 247–258. (42) Pukrakek, W. W. Engineering Fundamentals of the Internal Combustion Engine; Simon and Schuster Co.: New York, 1997; pp 278. (43) Ajav, E. A.; Singh, B.; Bhattacharya, T. K. Biomass Bioenergy 1998, 15 (6), 493–502.

temperature burned region, which is non-uniform, and the formation rates are highest in the regions closest to the stoichiometric region.44 In this study, it was discovered that increasing the methanol ratio in the blend raised NOx emissions. For example, as presented in Figure 4b, the change in NOx emissions was compared to M0 and showed that NOx augmented by 13.58, 34.56, and 49.38% for M5, M10, and M15, correspondingly, at 15 Nm load and ORG injection timing. Methanol contains 34% oxygen, and its cetane number is lower than diesel fuel, which boost peak temperature in the cylinder. On the other hand, the lower heating value (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 temperature in the cylinder. However, as shown in Figure 6, the exhaust temperature increased with an increasing methanol ratio in the fuel mixture. It is clear from the figure that the cetane number and oxygen content are more effective than LHV and (44) Agarwal, A. K. Prog. Energy Combust. Sci. 2007, 33, 233–271.




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.45 Unlike CO and UHC emissions, NOx emissions increased with an increasing engine load. For example, in comparison to Figure 3c and 4c, it was detected that NOx emissions increased by 5.71% for M5 as the engine load augments from 10 to 15 Nm constant loads at the retarded injection timing. Parts a-c of Figure 4 demonstrate the percent change in NOx emissions with different methanol blends at different injection timings for 15 Nm load. As shown in the figures for M15, retarded injection timing reduced in NOx emission by 9.66% and advanced injection timing increased in NOx emission by 6.32% compared to ORG injection timing, correspondingly. When the injection timing was retarded, it was observed that NOx emissions

decreased for all fuel mixtures. Retarding the injection timing decreased the peak cylinder pressure because more of fuel burned after TDC. Lower peak cylinder pressures resulted in lower peak temperatures. As a consequence, the NOx concentration diminished.35 Smoke Number (SN). Soot is formed in the cylinder, from heavy hydrocarbons in the gas phase, which condense and coalesce in the oxygen-deficient regions in the very rich core of the fuel sprays.46 Regarding the effect of different methanol contents on SNs, it was observed that increasing methanol ratio in the blend reduced SNs. The change in SNs compared to M0 implied that they diminished by 14.89, 18.80, and 25.53% for M5, M10, and M15, respectively, at 20 Nm load and advance injection timing as illustrated in Figure 5a. The presence of atomic-bound oxygen in methanol satisfies positive chemical control over soot formation. The tendency to generate soot by

(45) Nwafor, O. M. I.; Rice, G.; Ogbonna, A. I. Renewable Energy 2000, 21, 433–444.

(46) Challen, B.; Baranescu, R. Diesel Engine Reference Book, 2nd ed.; Butterworth and Heinemann Publishing: Oxford, U.K., 1999; p 480.


Canakci et al.


the fuel dense region inside a diesel diffusion flame sheath is reduced, so that soot-free spray combustion could be achieved.33 The formation of smoke is most strongly dependent upon the engine load. As the load increases, more fuel is injected, and this increases the formation of smoke.39 The results obtained in this study supported this statement. For instance, in comparison to Figure 4c and 5c, it was seen that the change of SNs increased by 5.71% for M10 as the engine load increased from 5 to 15 Nm at retarded injection timing. Advancing the injection timing reduced the smoke emissions. The earlier injection led to higher temperatures during the expansion stroke and more time in which oxidation of the soot particles occurred.46 Parts a-c of Figure 3 present the percent change in SNs at different injection timings for 10 Nm load. As seen from these figures for M15, advanced injection timing lowered in SNs by 2.67% and retarded injection timing raised in SNs by 2.63% compared to ORG timing for M15, respectively. Carbon Dioxide (CO2) Emissions. CO2 emission is produced by complete combustion of fuel. As illustrated in the Figure 3a, when the methanol amount increased in the fuel mixture,

the percent change in CO and UHC decreased. The percent change in CO2 had an opposite behavior when compared to the CO concentrations, and this was due to improving the combustion process as a result of the oxygen content in the methanol. The maximum increase in the change of CO2 was observed at 24.19, 32.61, and 73.61% for M5, M10, and M15, respectively, compared to M0 at 20 Nm engine load and advanced injection timing. In this study, CO2 emissions increased with the advanced injection timing for all fuel mixtures. As shown parts a-c of Figure 4 for M10, advanced injection timing increased the change in CO2 by 5.79% and retarded injection timing diminished in CO2 by 7.55% compared to ORG and retarded injection timing, respectively. Combustion Analysis. Figures 7 and 8 show the cylinder gas pressure, and Figures 10 and 11 demonstrate the rate of heat release for different fuel blends and ORG injection timing at 20 and 10 Nm loads, respectively. Figures 9 and 12 illustrate the cylinder gas pressure and rate of heat release for M0 and M15 at different injection timing and 20 Nm load, correspond-




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

ingly. The measured start of combustion and ignition delay for each fuel is shown in Table 6 at different injection timing. Peak Cylinder Gas Pressure. To analyze the cylinder gas pressure, the pressure data of 100 cycles with a resolution 0.75 °CA was averaged. Figure 7 shows the cylinder gas pressure with respect to the crank angle at 20 Nm load and ORG injection timing. As seen in the figure, peak cylinder gas pressure slightly decreased with the increase of the methanol supplement rate. The peak cylinder pressure occurred at 7.96 (at 3.20 °CA ATDC), 7.86 (at 3.28 °CA ATDC), 7.78 (at 3.32 °CA ATDC), and 7.77 MPa (at 3.44 °CA ATDC) for M0, M5, M10, and M15 at 20 Nm load and ORG injection timing, respectively. Lowering the cetane number by methanol addition was responsible for the increase in the ignition delay. The increase in the ignition delay would burn more fuel in the premixed burning phase. Because of this, the rate of pressure rise increased and peak cylinder gas pressure diminished.21 When Figure 7 is compared to Figure 8, it is seen that the cylinder gas pressure increased with an increasing engine load. Experimental results show that the increase in the cylinder


Canakci et al.

Figure 7. Cylinder gas pressure versus CA at 20 Nm load and ORG injection timing.

Figure 8. Cylinder gas pressure versus CA at 10 Nm load and ORG injection timing.

pressure was approximately 5.45% for M10 when the engine load increased from 10 to 20 Nm. The fuel consumption per unit time was measured between 0.42 and 0.81 g/s at these test conditions (10 and 20 Nm load at ORG injection timing for M10, respectively). Because of increasing fuel consumption per unit time with increasing engine load, this behavior provided an increase in the maximum cylinder gas pressure. The locations of maximum cylinder gas pressure approached TDC at 20 Nm engine load because starting the fuel injection occurred earlier than that of 10 Nm. Figure 9 shows a comparison of the changes in the cylinder gas pressures with respect to the crank angle obtained for M0 and M15 at different injection timings and 20 Nm load. The peak cylinder gas pressure was obtained at 8.47 (at 1.68 °CA ATDC), 7.96 (at 3.20 °CA ATDC), and 7.18 MPa (at 5.08 °CA ATDC) for advanced ORG and retarded injection timings, respectively. The trend was such that, as injection started earlier, peak pressures became higher, which applied to all fuel blends.

Also, the peak pressures occurred earlier with advancing injection timings.42,47 Rate of Heat Release (ROHR). As illustrated in the Figure 10, ROHR decreased with the increase of the methanol amount in the fuel blend. The maximum ROHR was obtained at 31.13 (at 8.93 °CA BTDC), 29.86 (at 8.90 °CA BTDC), 27.33 (at 8.86 °CA BTDC), and 26.95 (at 8.85 °CA BTDC) kJ/deg for M0, M5, M10, and M15 at 20 Nm load and ORG injection timing, respectively. Because methanol does not evaporate as easily as diesel fuel, the ignition delay increases with an 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 an increase in the ROHR. Conversely, the LHV of methanol is lower than diesel fuel, which reduces ROHR. It is apparent from the figure that the LHV of (47) Ozsezen, A. N.; Canakci, M.; Sayin, C. Energy Fuels 2008, 22, 1297–1305.



Figure 9. Cylinder gas pressure versus CA at different injection timing and 20 Nm load.

Figure 10. Rate of heat release versus CA at 20 Nm load and ORG injection timing.

methanol was more significant than the heat of vaporization with regard to ROHR.26 Figures 10 and 11 show the ROHR for different fuel blends and ORG injection timing at 20 and 10 Nm loads, respectively. As seen in the figures, ROHR increased with the rise in the 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 the crank angle obtained for M0 and M15 at different injection timings and 20 Nm load. The ignition delay in a diesel engine is defined as the time between the start of fuel injection and the start of combustion. As seen in Table 6, the ignition delay was raised with advanced injection timing for M0 and M15 because the fuel was injected earlier in the combustion chamber. This led to greater accumulation of the fuel in the ignition delay period and an increasing premixed heat release. This is the reason of increasing ROHR. Retarding injection timing led to a lower accumulation of fuel and poor combustion. However, both physical and chemical reactions must take place before a

significant fraction of chemical energy in the fuel is released. These reactions need a finite time to occur. However, as ignition delay proceeds, the in-cylinder pressures and temperatures decrease and reduce the favorable conditions for ignition. The most favorable timing for ignition lies in between these two conditions.48 Conclusions In this study, the influence of injection timing on the combustion characteristics, engine performance, and exhaust emissions of a diesel engine have been experimentally investigated using methanol-blended diesel fuel. The results indicated (48) Kumar, M. S.; Ramesh, A.; Nagalinman, B. Biomass Bioenergy 2003, 25, 309–318. (49) MERCK. Product Specification, Germany, 2006. (50) TUPRAS. Product Specification, Turkey, 2005 (in Turkish). (51) Lombardini. Engine Technical Specification, Turkey, 2000 (in Turkish).


Canakci et al.

Figure 11. Rate of heat release versus CA at 10 Nm load and ORG injection timing.

Figure 12. Rate of heat release versus CA at different injection timing and 20 Nm load.

that NOx emissions increased, smoke number and CO and unburned HC emissions decreased by methanol addition, and CO2 emissions increased because of the improved combustion. Using methanol-blended diesel fuels, smoke number and CO and unburned HC emissions reduced 7-32, 15-85, and 13-48%, while CO2 and NOx emissions increased 8-22 and 6-74%, respectively, depending upon the engine test conditions. Increasing the amount of methanol in the fuel mixture produced higher peak temperature in the cylinder. This effect increased NOx emissions. The results confirmed that increasing the ratio of methanol in the fuel blend leads to a decrease in the BSFC, BSEC, and BTE. This is probably of the result of the LHV of methanol, which is distinctly lower than that of the diesel fuel. The peak cylinder pressure and ROHR decreased with an increasing methanol ratio in the blend. A lower cetane number with methanol supplement leads to the increase in the ignition delay. Increasing the ignition delay caused the deteriorating combustion, and the peak cylinder pressure decreased. The increase in

the ignition delay increased the rate of pressure rise and reduced the peak cylinder gas pressure. In terms of injection timing, the test results demonstrated that, with advancing the injection timing, smoke number and CO and unburned HC emissions decreased while NOx and CO2 emissions increased. When the injection timing was advanced, CO emission decreased because of the improving reaction between fuel and oxygen. This caused an increase in the CO2 emissions. Advancing the injection timing caused an earlier start of combustion relative to the TDC. Because of this, the cylinder charge, being compressed as the piston moved to the TDC, had relatively higher temperatures and, thus, lowered the smoke number and UHC emissions and increased NOx emissions. With the advancing injection timing, the best results were obtained for the UHC and CO emissions at 20 Nm and the best results were gained for the smoke number at 5 Nm load. At these conditions, the smoke number and CO and UHC emissions were found as 62%, 0.06%, and 12 ppm, for M15, respectively. On the other hand, retarding the injection timing at 5 Nm load



Table 6. Combustion Characteristics of the Fuels at Different Injection Timing M0 20 °CA BTDC (ORG injection start of injection (°BTDC) 20 start of combustion (°BTDC) 9.27 ignition delay (°) 10.73

M10

M15

timing), 20 Nm 20 20 8.92 8.34 11.08 11.16

M5

20 8.01 11.99

15 °CA BTDC, 20 Nm start of injection (°BTDC) 15 15 start of combustion (°BTDC) 5.98 5.67 ignition delay (deg) 9.02 9.33

15 5.34 9.66

15 5.01 9.99


25 10.51 14.49

25 9.12 15.88

Injection Timing), 10 Nm 20 20 20 5.97 5.54 5.24 14.03 14.46 14.76

20 4.91 15.09

20 °CA BTDC (ORG start of injection (°BTDC) start of combustion (°BTDC) ignition delay (deg)


15 2.04 12.96

15 1.52 13.48


25 6.76 18.24

25 5.37 19.63

presented the minimum results of NOx and CO2 emissions. At these working conditions, NOx and CO2 emissions were measured as 5 ppm and 5.45%, respectively. The ORG injection timing gave the best results for BSFC, BSEC, and BTE compared to the other injection timings. When

the injection timing was advanced, the ignition delay was longer and this led to a reduction in engine output power. Thus, fuel consumption per output power was increased. On the other hand, retarded injection timing diminished the peak cylinder pressure because more fuel was burned at near TDC, and this result increased fuel consumption per output power. Also, advanced injection timing boosted peak cylinder pressure and ROHR because of the increase in ignition delay. Acknowledgment. This study was supported by the grant from the Scientific Research Project Commission of Marmara University (Project BSE-075/131102).

Nomenclature ATDC ) after top dead center BTDC ) before top dead center BTE ) brake thermal efficiency (%) BSEC ) brake-specific energy consumption (MJ/kWh) BSFC ) brake-specific fuel consumption (g/kWh) CA ) crank angle CI ) compression ignition CO ) carbon monoxide CO2 ) carbon dioxide UHC ) unburned hydrocarbon LHV ) lower heating value (kJ/kg) (LHV)b ) lower heating value of a given component in the fuel blend (kJ/kg) NOx ) nitrogen oxides ppm ) particulate per million rpm ) revolution per minute TDC ) top dead center EF800398R

Exhaust Emissions and Combustion Characteristics of a Direct Injection

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