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Experimental Investigation of Combustion Characteristics and Emissions of an Indirect Injection Diesel Engine under Different Operating Conditions† A. Turkcan and M. Canakci* Department of Mechanical Education, Kocaeli UniVersity, 41380 Kocaeli, Turkey, and AlternatiVe Fuels Research and DeVelopment Center, Kocaeli UniVersity, 41040 Kocaeli, Turkey ReceiVed May 12, 2008. ReVised Manuscript ReceiVed July 10, 2008
In this study, the effects of engine-operating conditions on the combustion characteristics and emissions of an indirect injection (IDI) diesel engine have been experimentally investigated. The engine tests were conducted on 20, 30, and 40 N m constant loads and at full load condition. The experimental results show that the maximum cylinder gas pressures (Pmax) increased with an increasing engine speed and load. The location of the maximum cylinder gas pressure closed to top dead center (TDC) at high engine load. The starts of combustion (SOC) points closed to TDC at high engine speeds. The ignition delay (ID) slightly decreased with the increase of the engine loads, and it increased with an increasing engine speed. However, as the total combustion duration (TC) increased with an increasing engine load, it decreased with high engine speeds. The exhaust emissions, such as nitrogen oxides (NOx), smoke opacity, unburned hydrocarbon (HC), and carbon monoxide (CO), were also discussed for different operating conditions.
Table 1. Engine Specification engine combustion type number of cylinder bore × stroke compression ratio injection pump injector opening pressure nozzle hole diameter maximum power
1.8 VD diesel BMC indirect injection and naturally aspirated water cooled, four strokes 4 80.26 × 88.9 mm 21.47:1 mechanically regulated distributor type 130 bar 0.2 mm 38.8 kW at 4250 rpm
1. Introduction Combustion characteristics of an engine are very important information to interpret engine performance and exhaust emissions. They are also useful for engine design and optimization. Besides, the combustion characteristics, such as maximum cylinder gas pressure, the heat-release rate can be used to explain the effects of engine-operating conditions on the engine performance or to compare the alternative fuels under the same operating conditions.1-3 The combustion characteristics of the indirect injection (IDI) diesel engines are different from the direct injection (DI) diesel engines, because of greater heat-transfer losses in the swirl chamber.4 This handicap causes the brake-specific fuel consumption (bsfc) of the IDI engine to increase and the total engine efficiency to decrease compared to that of a DI diesel engine. † From the Conference on Fuels and Combustion in Engines. * To whom correspondence should be addressed. Telephone: +90262-3032285. Fax: +90-262-3032203. E-mail: mustafacanakci@ hotmail.com. (1) Canakci, M. Bioresour. Technol. 2007, 98, 1167–1175. (2) Ozsezen, A. N.; Canakci, M.; Sayin, C. Energy Fuels 2008, 22, 1297– 1305. (3) Ghojel, J.; Honnery, D. Appl. Therm. Eng. 2005, 25, 2072–2085. (4) Heywood, J. B. Internal Combustion Engine Fundamentals; McGrawHill International Editions: New York, 1988; pp 491-497.
Because of these disadvantages of the IDI diesel engines, most engine research has focused on the DI diesel engines. However, IDI diesel engines have a simple fuel injection system and lower injection pressure level because of higher air velocity and rapidly occurring air-fuel mixture formation in both combustion chambers of the IDI diesel engines.5 In addition, they do not depend upon the fuel quality6 and produce lower exhaust emissions7 than DI diesel engines. Especially, unburned hydrocarbon (HC) and carbon monoxide (CO) emissions were significantly lower in these engines, which have homogeny charge condition.8,9 The IDI diesel engines use the heat of the piston recess wall to vaporize the fuel. If the air flow in the combustion chamber is properly adapted, an extremely homogeneous air-fuel mixture with a long combustion period, low pressure increase, and therefore, quite combustion and higher oxidation can be achieved.5 Some researchers10,11 have explained that the NOx emissions increase with the increasing engine load in the DI diesel engines. However, the NOx emissions are higher at mid loads compared to the full load condition in the IDI diesel engines. On the other hand, the smoke emission of the IDI engine is higher than that of the DI engine.12 Rakopoulos et (5) Bosch Handbook. Diesel-Engine Management: An OVerView; Robert Bosch GmbH: Stuttgart, Germany, 2003; pp 24-27. (6) Owen, K.; Coley, T. AutomotiVe Fuels Reference Book; SAE: Warrendale, PA, 1995; p 375. (7) Abdel-Rahman, A. A. Int. J. Energy Res. 1998, 22 (6), 483–513. (8) Iwazaki, K.; Amagai, K.; Arai, M. Energy 2005, 30, 447–459. (9) Rakopoulos, C. D.; Antonopoulos, K. A.; Rakopoulos, D. C.; Giakoumis, E. G. Appl. Therm. Eng. 2006, 26 (14-15), 1611–1620. (10) Ferguson, C. R. Internal Combustion Engines, Applied Thermosciences; John Wiley and Sons: New York, 1986; pp 349-417. (11) Fujimoto, H.; Oura, S.; Morinaga, S.; Hashimoto, Y.; Senda, J.; Yamashita, T. SAE Paper 1999-01-3652, 1999. (12) Lander, T. J. A.; Leskinen, A. P.; Rantanen, L.; Raunemaa, T. M. EnViron. Sci. Technol. 2004, 38 (9), 2707–2714.
10.1021/ef800335w CCC: $40.75 2009 American Chemical Society Published on Web 08/20/2008
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Figure 1. Schematic diagram of the experimental setup. Table 2. Specifications of the Diesel Fuel24 property
diesel fuel
density at 15 °C) kinematic viscosity (mm2/s, at 40 °C) flash point (°C) cold filter plugging point (°C) carbon residue (% mass) cetane number total contamination (mg/kg) copper strip corrosion (3 h, at 50 °C) ash (% mass)
820-860 2.0-4.5 55 -15 maximum of 0.30 minimum of 46 maximum of 24 number 1 maximum of 0.01
(kg/m3,
Table 3. Pressure Transducer Technical Data pressure sensor (for cylinder gas pressure) specifications sensor type measuring range sensitivity natural frequency linearity
(bar) (pC/bar) (kHz) (% FSO)
piezoelectric 0-250 -26.09 90 < (0.5%
pressure sensor (for fuel line pressure) specifications sensor type measuring range sensitivity natural frequency linearity
(bar) (pC/bar) (kHz) (% FSO)
piezoelectric 0-150 -11.11 100 < (0.6%
in many studies.15-21 Rakopoluos et al.9,22 investigated the heatrelease rate of a turbocharged IDI engine having a very small prechamber and a very narrow connecting passageway for both chambers. Their results showed that the magnitude heat release and cumulative heat-release rate of the main chamber increased with an increase in the engine load. They obtained that the heatrelease rate improved faster with an increasing engine speed. Bowden et al.23 studied that the heat-release rates of the two types of IDI engines. They observed the height of the first peak and second hump of the heat-release rate. They reported that the first peak increases with cumulative heat input during the delay period and the second hump increases with the increasing load and becomes very small at light loads. As different from the studies given above, in this study, the maximum cylinder gas pressure and its location points, the maximum rate of pressure rise, the heat-release rate and cumulative heat-release rate, the start of combustion and injection points, ignition delay, and total combustion duration were comprehensively investigated and discussed at wide range speeds and loads. At the same time, the emissions in terms of CO, HC, NOx, and smoke opacity of the IDI engine were also discussed. 2. Experimental Section
al.13 reported that total unburned HC emissions were unaffected by the engine load and speed, and the CO emissions increased with an increasing engine load for an IDI engine. They obtained different results for NOx emissions at different loads and speeds. Parlak et al.14 observed that NOx emissions depend upon the engine speed and load in the IDI diesel engine. They determined that, when the injection timing is kept constant, the NOx emissions decreased with an increasing engine speed and load. The effects of different engine operating conditions on the combustion characteristics have been observed and performed
(13) Rakopolous, C. D.; Hountalas, D. T.; Rakopoulos, D. C. Energy Fuels 2004, 18, 1302–1308. (14) Parlak, A.; Yasar, H.; Hasimoglu, C.; Kolip, A. Appl. Therm. Eng. 2005, 25, 3045–3052.
In this study, a naturally aspirated, water-cooled, IDI diesel engine was used as a test engine. The engine specifications are shown in Table 1. A hydraulic dynamometer has been directly coupled to the engine output shaft. Figure 1 shows the schematic diagram of the experimental setup. The following parameters were (15) Huang, Z.; Lu, H.; Jiang, D.; Zeng, K.; Liu, B.; Zhang, J.; Wang, X. Bioresour. Technol. 2004, 95, 331–341. (16) Sato, H.; Sato, Y.; Ishida, S. SAE Paper 930720, 1993. (17) Selim, M. Y. E.; Radwan, M. S.; Elfeky, S. M. S. Renewable Energy 2003, 28, 1401–1420. (18) Benajes, J.; Molina, S.; Garci, J. M.; Riesco, J. M. Proc. Inst. Mech. Eng., Part D: J. Auto. Eng. 2004, 218, 1141–1148. (19) Neame, G. R.; Wallace, J. S. SAE Paper 930935, 1993. (20) Bolt, J. A.; Henein, N. A. Proc. Inst. Mech. Eng., Part: 3J 19691970, 184, 130–136. (21) Hotta, Y.; Nakakita, K.; Inayoshi, M.; Ogawa, T.; Sato, T.; Yamada, M. JSAE ReV. 1997, 18, 19–31. (22) Hountalas, D. T.; Rakopolous, C. D.; Rakopolous, D. C.; Giakoumis, E. G. SAE Paper 2005-01-0926, 2005. (23) Bowden, C. M.; Samaga, B. S.; Lyn, W. T. Proc. Inst. Mech. Eng., Part: 3J 1969-1970, 184, 122–130.
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Turkcan and Canakci Table 4. Exhaust Gas Analyzers and Their Accuracies
emission devices Kane-May Quintox KM9106 Bilsa MOD 500 Bosch RTM 430 smoke opacity tester
measuring values
technology
accuracy
NOx
electrochemical
CO, CO2, total unburned HC smoke opacity
infrared
recorded during the each test: engine speed, load, fuel consumption, air flow rate, and ambient, cooling water inlet-outlet, and oil and exhaust temperatures. No. 2 diesel fuel was used, and its properties are shown in Table 2. Fuel consumption was determined by weighing fuel used for a period of time on an electronic scale. Air consumption was measured using a sharp-edged orifice plate (ISO 5167) and an inclined manometer as shown in Figure 1. Six different digital thermocouples monitored the temperatures of the intake air, fuel, engine oil, exhaust gas, and coolant inlet and outlet. In this study, piezoelectric-type sensors were used to obtain cylinder gas pressure and fuel line pressure data. The technical specifications of the sensors are given in Table 3. Three different gas analyzers were used to measure the exhaust gas concentrations. Table 4 gives the information about the exhaust gas analyzers and their accuracies. The cylinder gas pressure sensor was installed on the first cylinder of the engine head. The cylinder gas pressure was obtained using a Kistler water-cooled piezoelectric sensor type 6061B. An AVL quartz pressure sensor 8QP500c was mounted on the fuel line of the first cylinder to measure the fuel line pressure. The outputs of the pressure sensors were amplified by a Kistler charge amplifier 5015A type. The output of the charge amplifier and a signal from the magnetic pick-up were converted to digital signals and recorded by an Advantech PCI 1716A data acquisition card, which has a 16-bit converter and 250 kS/s sample rate. The pressure and crank angle data were stored in a computer. A computer program was written to collect the pressure data, with a resolution of 0.25° of crankshaft angle. To analyze the cylinder gas pressure, a combustion analysis program was written. To eliminate cycle-cycle variation, the cylinder gas pressure data of 50 cycles were averaged using a computer program. Then, the pressure data was used to calculate the heat-release rate. Experiments had been performed after the test engine reached the steady-state conditions. The steady-state conditions were determined with the engine oil temperature (∼70 °C). The test engine was run at least 5 min after the test engine
Bosch technology
(5 ppm < 100 ppm (5% > 100 ppm 0.001 vol %, 0.01 vol %, and 1 ppm, respectively 0.1% degree of opacity
was loaded, and then data was collected for each test. The test procedure was repeated 3 times to verify the each engine test condition, and the results were averaged.
3. Heat-Release Analysis In the literature, the heat-release combustion model developed by Krieger and Borman, generally, has been used for combustion analysis.25-27 In this study, the main combustion chamber and the precombustion chamber are combined into a single-zone thermodynamic model because Li et al.28 expressed that there is no significant difference if the precombustion chamber is combined with the main combustion chamber. The heat-release analysis was based on the changes of the cylinder gas pressure and cylinder volume during the cycle. Therefore, some assumptions were made to calculate the heat-release rate. It was assumed that no passage throttling losses exist between both chambers. Large temperature gradients, pressure waves, leakage through the piston rings, fuel vaporization, and charge mixtures were ignored. Hence, the intake and exhaust valves assumed to be closed. After using these assumptions, the heat-release rate is calculated using the following formula: ˙) Q
[ k -k 1 ]P dVdθ + [ k -1 1 ]V dPdθ
˙ is the combination of the heat-release rate, P is the where Q cylinder gas pressure, V is the cylinder volume, θ is the crank angle, and k is the ratio of specific heats. As shown in Figure 2, parameters of combustion characteristics are ignition delay, start of combustion, and combustion duration, which are obtained from the heat-release curve. The heat-release curve in a diesel engine examines ignition delay and total combustion duration. The ignition delay is defined as the time between the start of injection and the start of combustion. The start of injection time is usually determined when the injector needle lifts off its seat or the fuel line pressure reached the injector nozzle-opening pressure. The start of combustion is defined as the point where the heat-release rate turns from negative to zero. The total combustion duration is defined as the time from the start of combustion to the end of the heat release. 4. Results and Discussion In this study, the engine tests were performed at full load and 20, 30, and 40 N m constant loads. The engine speeds were selected between 1000 and 3000 rpm, with the increment of 500 rpm. The relationship between the combustion characteristics and engine-operating conditions are focused. The combus-
Figure 2. Definition of combustion characteristics.
(24) TUPRAS. Product Specification, Izmit, Turkey, 2006. (25) Challen, B.; Baranescu, R. Diesel Engine Reference Book, 2nd ed.; Butterworth Heinemann: Woburn, MA, 1999; p 97. (26) Pulkrabek, W. W. Engineering Fundamentals of the Internal Combustion Engine; Prentice-Hall: Upper Saddle River, NJ, 1997; pp 6869. (27) Borman, G. L.; Ragland, K W. Combustion Engineering; McGrawHill:New York; 1998; pp 234-240. (28) Li, J.; Zhou, L.; Pan, K.; Jiang, D.; Chae, J. SAE Paper 952055, 1995.
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Figure 3. Pmax, its location, maximum rate of pressure rise, and bmep for all engine-operating conditions.
Figure 4. Combustion and injection characteristics for 1000 rpm.
tion and injection characteristics were defined as the maximum cylinder gas pressure and its location, the maximum rate of pressure rise, the heat-release rate and cumulative heat-release rate, the start of combustion and injection points, ignition delay, and total combustion duration. The exhaust emissions obtained during the tests were also discussed for different operating conditions. Cylinder Gas Pressure. The maximum cylinder gas pressures values, their locations, and maximum rate of pressure rise in terms of the CA versus the engine speed are presented in Figure 3 for all test loads. Figure 4 shows the combustion parameters versus CA for different engine test loads at 1000 rpm. It was observed that the Pmax is affected by engineoperating conditions as load and speed. The results show that Pmax was between 7.30 and 8.00 MPa for 1000 rpm and increased with the increasing engine load as shown in Figure
4. This can be explained by the higher engine loads caused early injection as shown in the fuel line pressure curves in Figure 4. As shown in Figure 6, the fuel/air ratios were increased with a rising engine load because of extended fuel injection duration. This state provided an increase in the maximum cylinder gas pressure. The location of the maximum cylinder gas pressures are within 4.5° and 5 °CA after top dead center (ATDC) for 1000 rpm. As shown in Figure 5, the start of injection and start of combustion occurred earlier with increasing engine load. Therefore, the locations of maximum cylinder gas pressure points were closed to TDC. This behavior occurred at all engine speeds. The maximum rate of pressure rise was calculated as seen in Figure 3. The results show that the maximum rate of pressure rise decreased with an increasing engine speed until 3000 rpm. After that, it was observed that the maximum rate of pressure rise suddenly increased at 3000 rpm. The highest
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Turkcan and Canakci
Figure 5. SOC, SOI points, ID, and TC durations for all engine-operating conditions.
value of the maximum cylinder gas pressures and fuel line pressures were measured at 3000 rpm. This result may be influenced by the air turbulence and atomization rate, which was reached to the maximum level in the cylinder at this engine speed. This increase in the cylinder gas pressure caused the maximum rate of pressure rise to suddenly increase at 3000 rpm. Figure 3 shows the maximum rate of pressure rise increased with an increasing engine load because of the fuel/air ratio increasing with an increasing engine load. Heat-Release Rate. As shown in Figure 4, the change direction of the fuel line pressures is the same as the change direction of the heat-release curves. This showed the relationship between the start of fuel injection and heat release. At low engine loads, because of the start of combustion, the peak point of the heat-release curve is close to TDC. The heat-release profile has a slight negative dip during the ignition delay period, which is mainly heat loss from the cylinder during the fuel vaporizing phase, and it is more obvious at the full loads and high engine speeds. This state significantly affected, at full load test conditions, the magnitude of the peak point of heat release and was lower than those of constant load test conditions. Besides, the first peak point magnitude of heat release was related to the amount of combustible fuel in the ignition delay period. Because the amount of fuel consumption increases with the increasing engine load, the amount of combustible fuel also increases in the prechamber. This behavior caused the peak point of the heat release to increase with increasing engine loads. Similar trends for the heat-release curves were observed at all other engine speeds. It was observed that the area under the heat-release rate curves increased with an increasing engine load. The cumulative heat-release rates were calculated and illustrated also in Figure 4. The results show that the cumulative heatrelease rate increased with engine load. The same tendency of the cumulative heat-release rates was observed at all engine speeds. Start of Combustion (SOC) and Start of Injection (SOI). The starts of combustion and injection timing versus the engine speed are presented in Figure 5 for all test loads. Fuel line pressures increased with increasing engine speed, and SOI points
occurred earlier with increasing engine speeds and loads as illustrated in Figures 4 and 5. The earliest SOI points were found at 3000 rpm, compared to those of all of the other engine speeds. It was observed that SOC points were closer to TDC at 3000 rpm than those of 2000, 1500, and 1000 rpm as shown in Figure 5. This state can explain that the engine is operated at full load and high speed and more fuel needs to be injected. More fuel causes an increase in the evaporation heat and the SOC point at the high engine speed closer to TDC. It was observed that the starts of the combustion points were closed to TDC with decreasing engine load as given in Figure 4. Ignition Delay (ID) and Total Combustion (TC) Duration. The ID and TC duration in terms of the CA versus the engine speed are presented in Figure 5 for all test loads. The ID increases with increasing engine speeds as illustrated in Figure 5. The high engine speeds induce better atomization and evaporation in the air-fuel mixing, so that the ignition delay period becomes shorter. It was observed that the ignition delay decreased with increasing engine loads and fuel-air ratio. This behavior can be explained by the maximum cylinder gas pressure, and temperature increased with an increasing engine load and fuel/air ratio. It was found that, at the constant load tests, the ignition delays are shorter than those of full load tests. If the ignition delay is long in an IDI diesel engine, this means that most of the fuel is in its prechamber. This effect was interpreted as an increase in the cooling effect of the fuel and will cause a temperature drop of the cylinder gases and an increase in the ignition delay at full loads as shown in Figure 4. The TC durations were obtained between the 38° and 83 °CA for the engine tests applied in this study. The fuel injection duration increased with an increasing engine load. This increase in the fuel injection duration provided the TC duration, which increased with increasing engine loads as shown in Figure 5. Because of the increased ignition delay and decreased mixing controlled burning phase, the total combustion duration decreased with the increasing engine speed. Exhaust Emissions. The exhaust emissions obtained during this study are shown in Figure 6. The NOx emissions decreased
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Figure 6. Emissions, fuel/air ratio, and exhaust gas temperature for all engine-operating conditions.
with an increasing engine load, although the Pmax increased with an increasing engine load. This state can be explained by the decreasing amount of air that entered into the cylinder with the increasing engine load. On the other hand, the Pmax, exhaust gas temperatures, and fuel consumption increased with higher engine speed at constant load conditions. This situation caused the NOx emissions to increase with an increasing engine speed at constant load conditions. Different values for NOx emissions were obtained at the full load test conditions in terms of engine speed. In addition, humidity has a large influence on NOx emissions. For that reason, in this study, the relative humidity of air was measured for the NOx correction factor, which was calculated as assured in ref 29. The smoke opacity level increased with the engine load (see Figure 6). Especially, greater smoke opacity was observed at full load, because the ignition delay at full load was higher than that of the constant loads. Besides, the smoke opacity decreased with light to mid speeds because of the fuel-air mixing, which was enhanced by the increasing engine speed. The smoke opacity results show that the fuel-air mixing becomes poor at higher speeds because of the available time not being enough for a complete chemical reaction. The CO emissions are low in the diesel engine because of the lean mixture. It is known that CO emissions are most affected by the fuel/air ratio, fuel type, combustion chamber design, atomization rate, start of injection timing, engine load,
and speed. The most important among these parameters is the fuel/air ratio.30 The results show that the CO emissions were affected by the engine speed and load as shown in Figure 6. The CO emissions increased with engine speed and load because the fuel/air ratio increased with engine speed and load. The unburned HC emissions were obtained significiantly lower for all engine test speed and load conditions. The unburned HC emissions were slightly lower at high speeds because of shorter ignition delays. At full load conditions, the Pmax and exhaust gas temperatures were higher than those of constant loads. This state allows for more oxidation of unburned HC emissions; therefore, the unburned HC emissions reduced at full load test conditions. 5. Conclusions In this study, the combustion, injection, and emission characteristics of an IDI diesel engine running at the various speed and load conditions were investigated. The following conclusions can be drawn from this study: (1) The maximum cylinder gas pressure increases with an increasing engine load (29) Society of Automotive Engineers, Inc. SAE Handbook; SAE: Warrendale, PA, 2001; Vol. 1, pp 1304-1306. (30) Ozsezen, A. N.; Canakci, M.; Sayin, C. Energy Fuels 2008, 22, 2796–2804.
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and speed. (2) The location of the maximum cylinder gas pressure was closed to TDC with the increasing engine loads. This behavior was determined for all engine speeds. (3) The maximum rate of pressure rise decreased with an increasing engine speed and it increased with load. (4) The peak point of the heat release and the area under the heat-release rate curves increased with increasing engine load. The cumulative heatrelease rate increased with engine load for all engine speeds. (5) The injection pressures increased with engine speeds, and the starts of fuel injection occurred earlier with increasing engine speed and loads. (6) The starts of the combustion were closed to TDC with decreasing engine load, while they were closed to TDC with increasing engine speeds. This behavior was obviously observed at high engine speeds. (7) The ignition delay period increased with the increase in the engine speed. Although the ignition delay was seen shorter at the higher engine constant load, it was found longer at full engine load. Similar behaviors were observed for all engine speeds. (8) The total combustion
Turkcan and Canakci
duration increased with an increasing engine load because it decreased with an increasing engine speed. (9) The NOx emissions decreased with an increasing engine load and increased with engine speed. The different values for NOx emissions were obtained at full load test conditions in terms of engine speed. (10) The smoke opacity increased with an increasing engine load. However, it gave different results in terms of engine speed. (11) The CO emissions increased with engine speed and load. The unburned HC emissions were slightly lower at high speeds and reduced at full load test conditions. Acknowledgment. This study was supported by the Grant from TUBITAK (Project 104M372). The authors thank the individuals at the engine test laboratory, who were involved in making this work possible. EF800335W
