The influence of operating parameters on the performance and emissions of a DI diesel engine using methanol-blended-diesel fuel

Fuel 89 (2010) 1407–1414

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The influence of operating parameters on the performance and emissions of a DI diesel engine using methanol-blended-diesel fuel Cenk Sayin a, Ahmet Necati Ozsezen b,c, Mustafa Canakci b,c,* a

Department of Automotive Engineering Technology, Marmara University, 34722 Istanbul, Turkey Department of Automotive Engineering Technology, Kocaeli University, 41380 Izmit, Turkey c Alternative Fuels R&D Center, Kocaeli University, 41275 Izmit, Turkey b

a r t i c l e

i n f o

Article history: Received 8 September 2009 Received in revised form 22 October 2009 Accepted 28 October 2009 Available online 14 November 2009 Keywords: Alternative fuel Methanol Diesel engine Injection pressure Injection timing

a b s t r a c t In this study, the effects of injection pressure and timing on the performance and emission characteristics of a DI diesel engine using methanol (5%, 10% and 15%) blended-diesel fuel were investigated. The tests were conducted on three different injection pressures (180, 200 and 220 bar) and timings (15°, 20°, and 25° CA BTDC) at 20 Nm engine load and 2200 rpm. The results indicated that brake specific fuel consumption (BSFC), brake specific energy consumption (BSEC), and nitrogen oxides (NOx) emissions increased as brake thermal efficiency (BTE), smoke opacity, carbon monoxide (CO) and total unburned hydrocarbon (THC) decreased with increasing amount of methanol in the fuel mixture. The best results were achieved for BSFC, BSEC and BTE at the original injection pressure and timing. For the all test fuels, the increasing injection pressure and timing caused to decrease in the smoke opacity, CO, THC emissions while NOx emissions increase. Ó 2009 Elsevier Ltd. All rights reserved.

1. Introduction With the increasing concern about fuel shortage and environmental protection, the research on improving fuel economy and decreasing exhaust emissions has become the major activity in the engine combustion and development. Because of the limited reserves of crude oil, the development of alternative fuel engines has attracted increasing attention. Alternative fuels are usually clean compared with diesel and gasoline in the combustion processes. The introduction of these alternative fuels is beneficial to slow down the petroleum consumption and to reduce exhaust emissions [1]. Among these fuels, methanol, which can be used as either blend of the conventional fuels in the existing engines or an additive in biodiesel production, has received increasing attention. Because of its higher octane number and oxygen content, combustion of methanol in spark ignition engines shows better results compared to gasoline [2]. However, it has been reported that there have been some difficulties in the ignition of the air–fuel mixture with the use of methanol in the diesel engines due to mainly its low cetane number, high latent heat of vaporization and long ignition delay [3]. In order to overcome these problems and to use methanol in diesel engines, different methods such as

* Corresponding author. Address: Department of Automotive Engineering Technology, Kocaeli University, 41380 Izmit, Turkey. Tel.: +90 262 3032285; fax: +90 262 3032203. E-mail address: [email protected] (M. Canakci). 0016-2361/$ - see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2009.10.035

alcohol fumigation, dual injection, methanol–diesel fuel blend and methanol diesel fuel emulsion have been employed by the researchers [4,5]. Exhaust emissions from diesel engines fueled with conventional diesel fuel have caused environmental pollution and global problems. The European Commission has published some directives (2005/55/EC for Euro 4/5, etc.) to reduce exhaust emissions from light- and heavy-duty diesel engines [6]. Since using methanolblended-diesel fuel can ease off the air pollution, many researchers have devoted themselves to study the effect of this alternative fuel on the engine performance and exhaust emissions of diesel engine. Huang et al. [7], for example, used various blend rates of methanol–diesel fuels in the engine tests. The results pointed out that the increase of methanol content in the mixture decreases smoke, CO and THC emissions but increases NOx emissions. Yao et al. [8] investigated the effect of diesel–methanol compound combustion (DMCC) on the application of the system to two diesel engines. In the DMCC system, diesel fuel was used for engine starting and low load operation. From medium to high load, a fixed amount of diesel fuel was maintained while extra energy is acquired by injection method into the intake manifold to form a homogeneous methanol–air mixture. The system was tested on two 4-cylinder diesel engines: one naturally aspirated and the other turbocharged. In both cases, DMCC was found to decrease BSFC, NOx, and smoke emissions but to increase CO and THC emissions. Udayakumar et al. [9] studied the influence of methanol fumigation on the exhaust emissions. They carried out the tests

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with inlet air heated to 70 °C. They reported that smoke and NOx emissions reduce with methanol fumigation into the intake manifold. In another work [10], an electronically controlled low-pressure common rail system was employed to deliver methanol to inlet port, while the engine’s original high-pressure diesel injection system was used to deliver a suitable quantity of diesel fuel for ignition. The experimental results showed that the fuel-load power of the dual-fuel engine can reach or even exceed that of the original diesel engine when a suitable minimum pilot diesel quantity is used. Under the dual-fuel conditions, smoke was decreased significantly, while a modest reduction in NOx was observed. The BSFC was improved under the high-load operating conditions. However, THC and CO emissions for dual-fuel operation increased when methanol was added. The fuel injection systems play an important role to reduce the emissions to an acceptable level. Two of the most important injection characteristics are injection pressure and timing. Various investigators have reported that injection pressure has an effect on the engine performance and exhaust emissions. Can et al. [11] investigated the influence of ethanol ajddition (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. The results demonstrated that increasing the injection pressure of the engine running with ethanol–diesel decreased CO and smoke opacity while it increased NOx emission. Puhan et al. [12] investigated the effect of injection pressures on the performance and emissions characteristics of a diesel engine. In that study, a high linolenic linseed methyl ester was investigated in a diesel engine with varied fuel injection pressures (200, 220 and 240 bar). The results showed that the optimum fuel injection pressure is 240 bar with linseed methyl ester. At this optimized pressure, the BSFC and BTE were similar to diesel fuel, while a reduction in smoke opacity, CO, and THC emissions and an increase in the NOx emissions were noticed compared to diesel fuel. Icingur and Altiparmak [13] worked on the influence of injection pressures and fuel cetane numbers on a diesel engine. For this aim, 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). They reported that increasing the injection pressure decreases the smoke opacity and increases the NOx and SO2 emissions. Several researchers have also reported that the injection timing affects the engine performance and exhaust emissions of diesel engines. Payri et al. [14] conducted a study on a diesel engine operating at light load with standard injection and combustion systems. They retarded the start of injection timing for the aim of promoting the first phase of combustion in premixed condition. They reported that retarded fuel injection produces very low levels in smoke opacity and NOx emissions, but it causes to higher CO and THC emissions and a significant penalty in BSFC. At the light engine load, the combination of retarded fuel injection and introduction of exhaust gas recirculation (EGR) were proved to be very efficient in achieving smoke opacity and NOx levels. Aktas and Sekmen [15] investigated the effects of fuel injection advance on the performance and exhaust emissions of a fourstroke, single cylinder DI diesel engine fueled with biodiesel. Engine torque, brake power, BSFC, exhaust gas temperature, CO, THC, and NOx emissions were measured for injection timings of 24.9°, 26.6° and 28.5° crank angle (CA) at full load. When injection timing was increased to 26.6° CA, the engine torque increased by 6% and BSFC improved by 8%. In addition, it was also determined that CO and THC emissions decreased, while NOx emissions increased between 4% and 11%. Sayin and Canakci [16] studied the effect of injection timing on the exhaust emissions and engine performance of a DI diesel en-

gine using ethanol-blended-diesel fuel. The original start of injection timing of the engine was 27° CA BTDC in that study. The tests were conducted at five different injection timings; 21°, 24°, 27°, 30° and 33° CA BTDC. The results showed that BSFC, emissions of NOx and CO2 increased as BTE, emissions of CO and THC decreased with 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 THC and CO emissions decreased at the retarded injection timings (21° and 24° CA BTDC) in the all test conditions. On the other hand, with the advanced injection timings (30° and 33° CA BTDC), THC 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 when compared to the results of original injection timing. Based on the above considerations, the objective of this study is to clarify the effects of injection pressure and timing on the performance and exhaust emissions of a DI diesel engine using methanol-blended-diesel fuel. 2. Experimental setup and procedure The experiments were conducted using a single cylinder, fourstroke DI diesel engine. The main specifications of the test engine are given in Table 1 and set up of the test bench is shown in Fig. 1. In order to determine the engine torque, the shaft of the test

Table 1 The specifications of the test engine. Model of engine Type Cylinder number Bore Stroke Total cylinder volume Injector opening pressure Number of nozzle hole Nozzle hole diameter and number Start of injection timing Compression ratio Maximum torque Maximum power

Lombardini 6 LD 400 Direct injection, naturally aspirated, and four-stroke 1 86 mm 68 mm 395 cm3 200 bar 4 0.2 mm and 4 20° CA BTDC 18:1 21 Nm at 2200 rpm 7.5 kW at 3600 rpm

Fuel consumption measuring device Fuel line pressure sensor Air consumption measuring device Signal conditioner Exhaust emission control devices

Data acquisition board

K Type thermocouples

Angular referance

Dynamometer

Fig. 1. Experimental set-up.

C. Sayin et al. / Fuel 89 (2010) 1407–1414

engine was coupled to an electrical dynamometer, which was loaded by electrical resistance. A strain load sensor was employed to determine the load on the dynamometer. The engine speed was measured by electromagnetic speed sensor installed on dynamometer. Fuel consumption was quantified by combined container method. Air consumption was measured by inclined manometer. The engine was equipped with an orifice meter connected to an inclined manometer to measure mass flow rate of the intake air. An air damping was used for damping out the pulsations produced by the engine. The temperatures of air inlet, exhaust gas engine outlet, and engine oil were measured by K type thermocouples. CO, CO2 and THC emissions were measured with an infrared gas analyzer (Bilsa Mod 210) with an accuracy ±0.01%, ±0.01%, and ±1 ppm, respectively. Smoke opacity was measured with the use of a Bosch system with an accuracy ±0.1%. NOx emissions were recorded using an electrochemical gas analyzer (Kane-May Qintox KM9106) with an accuracy ±1 ppm. The analyzers were calibrated before the experiments. The accuracies of the measurements and the uncertainties in the calculated results are given in Table 2. Three fuel blends were prepared and employed in the experiments along with pure diesel fuel. The blends are 95% diesel fuel and 5% methanol (M5), 90% diesel and 10% methanol (M10) and 85% diesel and 15% methanol (M15) in volume. The fuel blends were prepared just before starting the each experiment to ensure that the fuel mixture is homogenous. A mixer was also used in the fuel tank in order to prevent phase separation. Thus, phase separation was not observed during the experiments. Diesel fuel was obtained from TUPRAS (Turkish Petroleum Refineries Corporation). Methanol, with a purity of 99%, was purchased from a commercial supplier. The fuel properties are shown in Table 3. The experiments were performed at two different modes. The original (ORG) injection pressure and timing of the engine are Table 2 The accuracies of the measurements and the uncertainties in the calculated results. Measurements

Accuracy

Load Speed Time Temperatures

±2 N ±25 rpm ±0.5% ±1 °C

Calculated results

Uncertainty

Power BSFC

±2.55% ±2.60%

Table 3 Properties of the fuels used in the tests.

Chemical formula Molecular weight (g/mol) Flame spread rate (m/s) Flame temperature (°C) Boiling temperature (°C) Density (g/cm3, at 20 °C) Flash point (°C) Auto-ignition temperature (°C) Lower heating value (MJ/kg) Cetane number Octane number C/H ratio Viscosity (mm2/s, at 25 °C)) Carbon content (wt.%) Hydrogen content (wt.%) Oxygen content (wt.%) Sulfur content (ppm) Stoichiometric air–fuel ratio Heat of vaporization (MJ/kg)

Methanol

Euro Diesel

CH3OH 32 – 1890 64.7 0.79 11 316 20.27 4 110 0.25 0.59 37.5 12.5 50 – 6.66 1.11

C14H28 196 2–4 2054 190–280 0.84 78 464 42.74 56.5 Not applicable 0.50 3.35 86 14 0 <50 14.28 0.27

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200 bar and 20° CA BTDC, respectively. Washer, located in the connection place between nozzle and injector spring, is 0.20 mm and adding one shim increases the injection pressure by 20 bar. In the first step, the experiments were carried out at three different injection pressures (180, 220 and 220 bar) at ORG injection timing. Thickness of one advance shim, located in connection place between engine and fuel pump, is 0.25 mm and adding or removing one shim changes the injection timing 5° CA. In the second step, the tests were performed in three different injection timings (15°, 20° and 25° CA BTDC) at ORG injection pressure. All tests were conducted at 20 Nm engine load and 2200 rpm. Each test was repeated three times and the results of the three repetitions were averaged. Tests firstly conducted with diesel fuel to obtain the base data of the engine. After each fuel test, the engine was run at least for 10 min to consume the fuel which left in the fuel system from the previous test. The values of the engine oil temperature, mass flow rate of air, exhaust temperature and pollutants such as CO, THC, NOx and smoke opacity, were recorded during the experiments. The ratios of the experimental values of the engine performance parameters and exhaust emissions obtained with the fuel blends to those with diesel fuel were determined and compared in graphics.

3. Results and discussions 3.1. Brake specific fuel consumption (BSFC) The BSFC is defined as the ratio of mass fuel consumption to the brake power. The percent variation in the BSFC with methanol–diesel fuel blends compared to diesel fuel is shown in Figs. 2 and 3. The obtained BSFC results for different injection pressures and timings are presented in Table 4. Due to low energy content of the methanol, there is an increasing trend in BSFC with increasing methanol content in the fuel mixture compared to diesel fuel. Among the tested fuels the lowest BSFC values were obtained with diesel fuel because of low fuel consumption rate. The results showed that BSFC increases significantly when the engine is fueled with the blends having high methanol content due to lower heating value and density of methanol. As shown in Table 4, BSFC values were acquired as 141.20 g/kWh with the M0; 149.23 g/kWh with the M5; 162.38 g/kWh with the M10; and 172.28 g/kWh with the M15, respectively at 180 bar. The minimum BSFC values were obtained at ORG injection pressure for all fuel blends. For instance, BSFC values were found as 120.84, 134.57 and 149.23 g/kWh at ORG, increased (220 bar) and decreased (180 bar) injection pressures for M5, respectively. This may be explained with the fuel particle diameters which will enlarge with decreasing injection pressure, and ignition delay period during the combustion will increase. This situation causes to increase in BSFC. On the other hand, Bakar et al. [17] stated that the increasing injection pressure causes shorter ignition delay period. So, with decreasing injection pressure, the possibilities of homogeneous mixing decrease and BSFC augments. As shown in Table 4, advanced or retarded injection timings caused to increase in BSFC values for all test fuels. For example, BSFC values were observed with 117.54 g/kWh at ORG injection timing, 134.56 g/kWh at advanced injection timing and 142.73 g/ kWh at retarded injection timing for M0. As explained by Raheman and Ghadge [18], the retarded injection timing means later combustion. Therefore, the cylinder pressure rises only when the volume expands rapidly and this reduces the effective pressure to the work. Thus, advancing the injection timing means that the combustion starts earlier and more fuel can burn before the piston reached to top dead center (TDC).

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a

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BSFC

BSEC

a

BTE

40

BSFC

BSEC

BTE

40 o

30

220 bar

25 CA BTDC

30

20

20

10

10 0

0

-10

-10

-20

-20

-30

-30 -40

-40

30

% Change in performance

b

40

20 10 0 -10 -20 -30

M5

M10

M15

c

40 30

o

20 CA BTDC (ORG injection timing)

20 10 0 -10 -20 M5

-30

M10

40 o

15 CA BTDC

30

180 bar

20 10

M15

-40

c

-40

40 30

200 bar (ORG in ection ressure)

% Change in performance

b

20 10 0 -10

0 -20

-10 -20 -30 -40

-30 -40

Fig. 3. The changes in the performance parameters with the blends compared to diesel fuel at different injection timings and ORG injection pressure.

Fig. 2. The changes in the performance parameters with the blends compared to diesel fuel at different injection pressures and ORG injection timing. Table 4 BSFC values at different injection pressures and timings.

3.2. Brake specific energy consumption (BSEC) The BSEC is described as multiplication of BSFC and LHV. The variation in the BSEC of the engine with methanol–diesel fuel blends compared to diesel fuel is shown in Figs. 2 and 3. As shown in the figures, the change in BSEC increases with the methanol content. Mainly, due to the lower energy content of the methanol–diesel fuel blends, their BSFC values are higher than that of diesel fuel as shown in Table 4. Other reasons for increases in the BSFC with methanol–diesel blends may be said as low density and viscosity of methanol as well as its low cetane number. The low fuel viscosity causes a high fuel leakage from the fuel system, thus yielding a higher BSFC. On the other hand, when cetane number of the fuel blend supplied to the engine decreases, the ignition delay increases. For these reasons, BSFC increases with the increasing methanol amount in the fuel blend. Thus, the engine consumes more fuel to maintain the same amount power output. As shown in Fig. 2b, the change in BSEC relative to diesel fuel is 2.6%, 5.4% and 9% for M5, M10 and M15, respectively, at ORG injection pressure. As observed in Fig. 2a–c, the minimum change in BSEC was attained at ORG injection pressure for the all fuel blends. When the

BSFC (g/kWh) Injection pressure (Bar)

M0 M5 M10 M15

Injection timing (°CA BTDC)

180

200 (ORG)

220

15

20 (ORG)

25

141.20 149.23 162.38 172.28

115.09 120.84 121.19 135.58

127.33 134.57 145.66 154.06

142.73 155.48 169.84 198.12

117.54 120.98 123.36 152.21

134.56 137.25 144.51 182.46

injection pressure was changed from ORG injection pressure, the percent change in BSEC 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 decreased injection pressures were 2.7% and 3.6% for M15, respectively. Advanced and retarded injection timings resulted in 5.2% and 9.3% increase in BSEC compared to the value of ORG injection timing for M15. It can be observed from Fig. 3a–c that the advanced and retarded injection timings are responsible for increasing BSFC, and also result in increasing BSEC.

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BTE of an engine is the efficiency in which the chemical energy of a fuel is turned into useful work. Compared to diesel fuel, the changes in the BTE of the engine using methanol–diesel fuel blends are shown in Figs. 2 and 3. There are very slight decreases in BTE with the use of M5 and M10 compared to diesel fuel. In the test operations with the fuel blends, the engine yields relatively high BTE values for the blends up to 10%, which can be attributed to the promoted combustion due to the oxygen content of the blends. Irshad [19] stated that the fuel blends up to 10% do not cause a significant decrease in the energy content and cetane number of the fuel. In BTE, a noticeable decrease was observed in the case of M15, which results from the considerable increase in BSFC as seen in Table 4. As shown in Fig. 3a, the change in BTE relative to M0 is 7.5%, 9.6% and 26.7% for M5, M10 and M15, respectively, at advanced injection timing. As seen in the Figs. 2 and 3, increased or decreased injection pressure and injection timing diminished the change in the BTE by increasing BSFC seen in previous section. The BSFC values shown in Table 4 also confirm this situation. The best results of the change in the BTE were obtained at ORG injection pressure and timing by 2.6% and 0.6% for M5.

a

CO

THC

NOx

SN

80 60

220 bar

40 20 0 -20 -40 -60 -80

b

80 60

% Change in emisssions

3.3. Brake thermal efficiency (BTE)

200 bar (ORG injection pressure)

40 20 0 -20 -40 -60

M5

3.4. Carbon monoxide (CO) emissions

M10

M15

-80

The CO emissions in the exhaust gases represent lost chemical energy that is not fully used in the engine. Generally, CO emission is affected by air–fuel equivalence ratio, fuel type, combustion chamber design, atomization rate, start of injection timing, injection pressure, engine load, and speed. The most important among these parameters is the air–fuel equivalence ratio [6]. The variation in the CO emissions of the engine is shown in Figs. 4 and 5 when methanol–diesel fuel blends are compared to pure diesel fuel. One reason for the reduction in CO emissions of the fuel blends may be lower C/H ratio of methanol, as seen in Table 3. Moreover, methanol is an oxygenated fuel and it leads to more complete combustion. Hence, these effects decrease CO emissions in the exhaust. As demonstrated in Fig. 4a, the reductions in CO emissions compared to the results of diesel fuel are 26.7%, 40% and 50% for M5, M10 and M15, respectively, at increased injection pressure. Fig. 4a–c show the changes in the CO emissions for different methanol-blended-diesel fuels at different injection pressures. As seen in the figure, it was concluded that the increased injection pressure reduced the CO emission by 6.3% and the decreased injection pressure increased the CO emission by 2.6% compared to the results of ORG injection pressure for M15. Park and Lee [20] state that the increasing injection pressure causes a good fuel–air mixing, and easy and complete combustion of the smaller droplets. These effects lead to decrease in CO emissions. Fig. 5a–c demonstrate the change in the CO emissions for different methanol blends at different injection timings. As shown in the figure, it was concluded that the advanced injection timing decreased the CO emission by 13% and the retarded injection timing increased the CO emission by 3% compared to the results of ORG injection timing for M10. The advanced injection timing produces higher cylinder temperature and improves oxidation process between carbon and oxygen molecules [21]. These lead to diminish of the percent change in CO emissions. 3.5. Total unburned hydrocarbon (THC) emissions Unburned hydrocarbon emissions consist of fuel that is completely unburned or only partially burned. THC emission is mostly due to the retention of unburned fuel in crevices in the cylinder [22]. Figs. 4 and 5 show the percent changes in the THC emission

c

80 60

180 bar

40 20 0 -20 -40 -60 -80

Fig. 4. The changes in the emissions with the blends compared to diesel fuel at different injection pressures and ORG injection timing.

of the engine using methanol–diesel fuel blends. As seen in the figures, the THC emission was gradually reduced when the methanol ratio increased in the fuel blend. Relating to the effect of different methanol contents on THC emission, it was found that the increasing methanol ratio in the fuel blend decreased THC emission. For instance, as presented in Fig. 4a – at the increased injection pressure, the THC emissions decreased by 19%, 35% and 48% for M5, M10 and M15, respectively. When methanol is added to the diesel fuel, it provides more oxygen for the combustion process and leads to the improving combustion. Alla et al. [23] stated that methanol molecules are polar and cannot be absorbed easily by the non-polar lubricating oil; and therefore methanol can lower the possibility of the production of THC emissions. . Fig. 4a–c illustrate the change in the THC emission results for different methanol-blended-diesel fuels at different injection pressure. As shown in the figure, the increased injection pressure reduced the THC emissions. For M5, the reduction in the emission of THC was 4.3%, when the injection pressure was increased from 180 to 220 bar. As mentioned in the previous section, the increasing injection pressure causes better fuel–air mixture in the combustion chamber; therefore, the THC emissions are obtained less than that of the low injection pressure. Fig. 5a–c demonstrate the change in the THC emissions for different methanol-blended-diesel fuels at different injection timings.

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a

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CO

THC

NOx

SN

80 60

o

25 CA BTDC

40 20 0 -20 -40 -60 -80

b

80 o

% Change in emissions

60

20 CA BTDC (ORG injection timing)

40 20 0 -20 -40 M5

-60

M10

M15

-80

c

80 60

o

15 CA BTDC

40 20 0 -20 -40 -60 -80

Fig. 5. The changes in the emissions with the blends compared to diesel fuel at different injection timings and ORG injection pressure.

As seen in the figures, the advanced injection timing caused to reduction in THC emission by 11% and the retarded injection timing increased the THC emission by 2.6% compared to ORG timing for M10. Advancing the injection timing causes earlier start of combustion relative to the TDC. Therefore, the cylinder charge, being compressed as the piston moves to the TDC, has relatively higher temperatures, and thus lowers the THC emissions. 3.6. Nitrogen oxides (NOx) emissions The oxides of nitrogen in the exhaust emissions contain nitric oxide (NO) and nitrogen dioxides (NO2). During the combustion in an engine, oxygen atoms dissociate at high temperatures and start the chain reaction which is the oxidation of nitrogen to NO and NO2. Therefore, the local temperature, concentrations of oxygen and nitrogen atoms, and the time available for the oxygen– nitrogen reactions are the important parameters for NOx formation [24,25]. The variation in the NOx emission of the engine is shown in Figs. 4 and 5. As seen in the figures, the increasing methanol ratio in the fuel blend increased NOx emissions. For example, as illustrated in Fig. 5c – at the retarded injection timing, NOx emissions boosted by 14.3%, 39% and 45% for M5, M10 and M15, respectively. Methanol contains 34% oxygen and its cetane number is lower than

diesel fuel, which increased the peak temperature in the cylinder. Song et al. [26] indicated that increased oxygen levels increased the maximum temperature during the combustion, which increased NOx emissions. However, LHV of methanol is nearly two times lower than diesel fuel and latent heat of vaporization of methanol is about four times greater than diesel fuel, which decreases peak temperature in the cylinder. As shown in Table 4, the exhaust temperature increased along with increasing methanol ratio in the fuel blend. It is clear from the figure that cetane number and oxygen content are more effective than LHV and latent heat of vaporization regarded with increasing peak temperature in the cylinder. Therefore, the concentration of NOx increased when the methanol content was increased in the fuel blend. As presented in the Fig. 4a–c, the increased injection pressure increased the NOx emissions by 4% and the reduced injection pressure decreased them by 3.8% for M10, when compared with ORG injection pressure. When the injection pressure was decreased, it was observed that NOx emissions diminished for the all fuel blends. Purushothamana and Nagarajan [27] reported that the increasing injection pressures decreased the particle diameter and caused to quick vaporization of the diesel–methanol fuel spray. However, the liquid fuel cannot penetrate deeply into the combustion chamber. Therefore, higher injection pressure initially generates faster combustion rates, resulting in higher temperatures. As an in consequence, NOx formation starts to increase. As shown in Table 5, the obtained exhaust gas temperatures confirm this statement. Fig. 5a–c illustrate the percent change in NOx emissions for different methanol blends at different injection timings. As demonstrated in the figures, the advanced injection timing increased the NOx emissions by 11% and the retarded injection timing reduced them by 2% for M5, compared to ORG injection timing. When the injection timing was retarded, the NOx emissions decreased for the all fuel blends. Retarding the injection timing decreases the peak cylinder pressure because more fuel burns after the piston reached to TDC. Therefore, lower peak cylinder pressures results in lower peak temperatures. As a result, the NOx concentration starts to diminish.

3.7. Smoke opacities 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. Besides, among the particulate matter components, soot is recognized as the main substance responsible for the smoke opacity, and therefore, opacimeters usually convert their opacity measurements into soot concentrations. Smoke opacity formation occurs at the extreme air deficiency. The air or oxygen deficiency can be seen locally in combustion chamber. It increases as the air–fuel ratio decreases. Soot is produced by oxygen deficient thermal cracking of long-chain molecules [28,29]. The variation in the smoke opacity of the engine is shown in Figs. 4 and 5. As shown in Fig. 5b – at

Table 5 Exhaust gas temperatures at different injection pressures and timings. Exhaust gas temperature (°C) Injection pressure (Bar)

M0 M5 M10 M15

Injection timing (°CA BTDC)

180

200 (ORG)

220

15

20 (ORG)

25

223 239 240 248

230 241 243 266

237 255 260 291

220 230 234 254

230 241 243 266

234 248 252 280

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ORG injection timing, the smoke opacities decreased by 12.5%, 14.6% and 17.7% for M5, M10 and M15, respectively, when compared to diesel fuel. Smoke in diesel engines sources from the pyrolysis of heavy hydrocarbons [30]. The application of methanol will reduce diesel quantity injected into the cylinder. Because of the longer ignition delay and the higher latent of vaporization of methanol, more diesel fuel will be operated, which reduces diesel pyrolysis. Otherwise, the presence of methanol will act to enhance soot oxidation because of the increased concentration of oxygen atoms. Fig. 4a–c show the percent change in smoke opacity at different injection pressures. As seen in the figures, the increased injection pressure lowered the smoke opacity by 1% and the decreased injection pressure raised the smoke opacity by 3.7% for M15, when compared to ORG injection pressure. Increasing the injection pressure decreases the smoke opacity. When the injection pressure is increased, fuel particle diameters will become smaller. Therefore, fuel–air mixture will become better through the ignition period, and so smoke opacity will be less [31]. Fig. 5a–c illustrate the percent change in smoke opacity for different injection timings. As seen in the figures, the advanced injection timing lowered the smoke opacity by 3.5% and the retarded injection timing raised it by 2.4% for M10, when compared to ORG timing. The earlier injection causes more time for soot oxidation and leads to higher temperatures during the expansion stroke, and so the smoke emission reduces [32].

4. Conclusion Environmental protection is an important issue for the future of the world. Because of the reducing amount of petroleum reserves and its rising price, alternative fuels are intensively investigated for the full or partial replacement with diesel fuel. Therefore, in this study, the effect of injection pressure and timing on the performance and exhaust emissions of a DI diesel engine has been experimentally investigated during the usage of methanol-blendeddiesel fuel. The tests were conducted at 20 Nm constant engine load and 2200 rpm engine speed. The following conclusions can be drawn from the present paper; (1) When methanol-blended-diesel fuel was used, BSFC and BSEC increased due to the lower energy content of methanol. Also, all fuel blends yielded a decreased BTE in proportional to the methanol amount in the blend. (2) Compared to diesel fuel, all fuel blends yielded a decrease in the smoke opacity, THC, and CO emissions. Increasing the amount of methanol in the fuel blend produced higher exhaust temperatures in the tailpipe. (3) Compared to ORG injection pressure, the BSFC and BSEC increased, and BTE decreased along with the increased or decreased injection pressure, when methanol–diesel fuel blends were used. Smoke opacity, THC, and CO emissions decreased and NOx emissions increased with the increased injection pressure for the all fuel blends. Increasing injection pressure is the reason for better fuel–air mixing and complete combustion. This situation leaded to a decrease in smoke opacity and CO emissions. (4) When the injection timing was advanced, smoke opacity and CO emission decreased due to improving the reaction between the fuel and oxygen. 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 moves to the TDC, had relatively higher temperatures and thus, lowered THC emissions and increased NOx emissions.

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Acknowledgments The authors would like to acknowledge Marmara University for providing support to this research project and Kocaeli University, Alternative Fuels R&D Center for the contribution of the experiments.

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