Energy Fuels 2009, 23, 4981–4987 Published on Web 08/26/2009
: DOI:10.1021/ef8011382
Experimental Test of a New Compressed Natural Gas Direct Injection Engine Masjuki H. Hassan, M. Abul Kalam,* and TM Indra Mahlia Department of Mechanical Engineering, University of Malaya, Malaysia
Ishak Aris and M. Khair Nizam Department of Electrical Engineering, University of Putra Malaysia, Malaysia
Shahrir Abdullah and Yusoff Ali Department of Mechanical Engineering, University of Kebangsaan, Malaysia Received December 30, 2008. Revised Manuscript Received July 25, 2009
This paper presents experimental test results of a new compressed natural gas direct injection (CNGDI) engine that has been developed by modifying a multicylinder gasoline port injection engine. The major modifications done are (1) a modification of the injection system to gas direct injection using new highpressure gas injectors, (2) a change of the compression ratio from 10 to 14 through modification of piston and cylinder head, (3) and new long edge spark plugs. The CNG pressure at the common rail is kept constant at 20 bar. The engine has been operated at wide open throttle (WOT) condition to compare all the results with the original base engine. Hence, two engines are tested in this investigation. The test was conducted on a computer-controlled dynamometer to measure brake power; specific fuel consumption; volumetric efficiency; and exhaust emissions such as carbon monoxide (CO), oxides of nitrogen (NOx), and unburned hydrocarbon (HC). The objective of this investigation is to compare the test results between a CNGDI engine and a gasoline engine. It was found that the CNGDI engine produces a 4% higher brake power at 6000 rpm as compared to the original gasoline-fueled engine. The average brake specific fuel consumption (BSFC) of the CNGDI engine was slightly lower (about 0.28%) than the base engine. The CNGDI engine reduces 50% NOx emission as compared to base engine. However, the CNGDI engine produces higher HC and CO emissions as compared to the base engine by 34 and 48%, respectively. The results of this experiment will be used to further improve the CNGDI engine as well as to develop a new CNGDI car. The detailed results including fuel mapping are presented with discussions.
NOx. With increasing compression ratio, dual-fuel engines produced combustion noise as described in ref 1. According to ref 2, the HCCI engine is not suitable for operation on compressed natural gas (CNG) due to its high octane number, high autoignition temperature, and rapid heat release. The high friction loss of the engine was a dominant factor for limiting the lean operating region to a narrow window between insufficient power and knocking. As a result, HCCI operation on pure CNG was considered as unsuccessful because of extremely high NOx, high cyclic variation, and low efficiency. PCCI engines have a lower thermal efficiency due to the need to avoid engine knock and the unavoidable throttling at the intake at partial load. At low equivalence ratio, unstable ignition occurs due to the cooling fuel-air mixture on the cylinder wall. Low equivalence ratio with early injection increases NOx emissions, and late injection decreases heat release rate.3 Some researchers such as Westport Innovations Inc. and Isuzu car company (Japan) have proposed and developed
Introduction Natural gas can be the next prime fuel for energy conservation machinery system such as internal combustion (IC) engines. The existing IC engines are not efficient when operated with natural gas fuel due to its different physiochemical properties from commercial diesel and gasoline fuels. However, worldwide automotive engine researchers are involved in improving fuel efficiency with natural gas to meet stringent emission limits. Researchers’ efforts are going on through diesel and gasoline cycle piston-compression engine systems. The diesel cycle systems involved are dual fuel, homogeneous charge compression ignition (HCCI), and premixed charge compression ignition (PCCI) and CNGDI (with diesel cycle) systems. Researchers in various countries have carried out experimental work using gaseous fuels as diesel engine fuel substitute in a dual-fuel mode of operation. It is found that many parameters, namely, load, speed, pilot fuel injection timing, pilot fuel mass, compression ratio, inlet manifold conditions, and type of gaseous fuel, play an important role in the performance of dual-fuel diesel engines. Overall it was noted that dual-fuel engines produced lower brake power and higher
(1) Sahoo, B. B.; Sahoo, N.; Saha, U. K., Renew. Sust. Energ. Rev. 2009, 13, 1151-1184. (2) http://vhosseini.googlepages.com/VHosseiniMDCheckel.pdf (Accessed 5/20/2009) (3) Ishiyama, T.; Kawanabe, H.; Ohashi, K.; Shioji, M.; Nakai, S., Int. J. Engine Res., 2005, 6, 443-451.
*To whom correspondence should be addressed. E-mail: Kalam@ um.edu.my; Telephone: þ603-79676863; Fax: þ603-79675317. r 2009 American Chemical Society
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: DOI:10.1021/ef8011382
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The gas was injected during the compression stroke. It is believed that a systematic study encompassing fuel systems, combustion chambers, control units, vehicle body, fuel storage, and refueling infrastructure are required to develop the actual car based on a CNGDI engine. The main objectives of this investigation are: (i) To experimentally investigate the performance and emissions characteristics of a newly developed compressed natural gas direct injection (CNGDI) engine at WOT. (ii) To compare CNGDI and gasoline engines. Experimental Setup and Procedure A schematic diagram of the experimental setup is shown in Figure 2. A total of two engines were tested, and their specifications are shown in Table 1. The CNGDI engine was developed through modification of a gasoline engine. The major modifications are: (1) increasing the compression ratio from 10 to 14 through modifying piston and cylinder heads (some metal is added to the cylinder head and the piston head is modified from flat to semiflat to increase swirl), (2) new spark plugs with a long edge were used, and (3) the fuel injection system was modified from MPI to DI system. The CNG injection pressure was 20 bar at the common rail. The temperature of CNG at the common rail was 16 °C. The injector was designed to inject CNG fuel into the cylinder. The injector was initially set with a spring preload of 38 N. The spring preload was then adjusted with (1N to trim the dynamic flow at 100 Hz with a 2.0 ms pulse width. The average stroke length, dynamic flow rate, and opening and closing times are 0.267 mm, 19.06 mg/shot, and 1.50 and 0.93 ms, respectively. An eddy current dynamometer with a maximum absorption power of 150 kW was used to maintain the load at different engine speeds. The dynamometer could be started, loaded, and monitored via remote operation of the control-instrumentation unit and data acquisition control system. The dynamometer was also equipped with a speed sensor, switches for low pressure and high temperature for cooling water, the drive shaft, water inlet valve, and load cell torque measurement unit. The air flow rate into the engine inlet manifold was measured by a hot-wire anemometer (accuracy ( 0.2%) which comes with the engine control unit. This air mass flow meter data was transferred to an analog input card through a signal cable of 0-5 V. Finally, the actual airflow into the engine was analyzed from the data logged (Cadet 12 engine controlled software) into the computer. A coriolis micro motion mass flow meter was used to measure CNG flow rate into engine. The water and lubricant temperatures were controlled at 80 and 90 °C, respectively. A Horiba exhaust gas analyzer was used to measure the emission concentrations for the CNGDI engine. This analyzer was interfaced with the main engine controlled software (CADET12), so that all the emission data and engine operating data can be logged at the same time for analysis. This analyzer consists of an individual module for each emission parameters and has a zero and span gas calibration facility. The measurement techniques of this analyzer are infrared for CO, CO2, and HC, and chemiluminescence for NOx emissions. The detailed working principle can be seen on the Horiba Web site (horiba.com). A professional lambda meter was used to measure the exhaust air-fuel ratio to be tuned up by the ECU. It accurately determines the exhaust gas mixture strength over a wide range of engine operating conditions with a fast response time.
Figure 1. Various combustion systems for CNG fuel.
CNGDI engines based on the diesel cycle combustion system.4-6 It was proposed4-6 that natural gas direct injection and shielded glow plug ignition with a hot surface system mounted on the cylinder head would improve engine efficiency. It was found that the CNGDI engine produced lower brake power, increases NOx with lower equivalence ratio, and at high throttle smoke and misfire were found. Many researchers reported1-6 that CNG fuel in diesel produced high cycle-bycycle variations, misfire at high throttle, low heat release rate at late injection, and high NOx at early injection with low equivalence ratio etc. In this investigation, a CNGDI engine (with spark ignition) has been proposed to replace gasoline fuel and combustion system. It is based on an Otto cycle with spark plug ignition. The following figure (Figure 1) shows various combustion systems for CNG fuel. From the figure, it can be explained that each combustion system has unique features to reflect specific strategies of mixture preparation, combustion control, and emission reduction. However, all systems have a common goal of achieving substantial fuel economy improvement while simultaneously achieving large reductions in engine output and tailpipe emissions. Huang et al.7-10 have studied direct injection systems using a rapid compression machine and a single-cylinder gasoline engine. A single-cylinder engine was modified into a direct injection engine. The injector was placed in the intake manifold to inject gas into the cylinder at 80 bar. The gas was injected at the intake manifold early during the compression stroke. Their study was helpful to understand ignition/injecting timings as well as combustion and emissions characteristics. In this investigation, a dedicated injector was designed and placed at the top of the engine cylinder (near spark plug) to inject natural gas into cylinder at a constant pressure of 20 bar. (4) Sandeep, M.; Patric, O.; James, H.; Costi, N.; Jeff, T.; Stewart, W. Direct Injection of Natural Gas in a Heavy-Duty Diesel Engine, SAE Paper No. 2002-01-1630. (5) Dale, G.; Mark, Dunn.; Sandeep, M.; Edward, L.; Wright, J.; Vinod, D.; Mike, F. Development of a Compression Ignition Heavy Duty Pilot-Ignited Natural Gas-Fuelled Engine for Low NOx Emissions, SAE Paper No: 2004-01-2954. (6) Frailey, M. R.; Lyford-Pike, E.; Kamel, M. M.; Wayne, S.; Nine, R. D.; Clark, N. N.; Bolin, M. S. An Emission and Performance Comparison of the Natural Gas Cummins Westport Inc. C-Gas Plus Versus Diesel in Heavy- Duty Trucks, SAE Paper No: 2002-01-2737. (7) Huang, Z.; Shiga, S.; Ueda, T.JSME Int. J. B: Fluids Therm. Eng. 2002, 45 (4), 891-900. (8) Huang, Z.; Shiga, S.; Ueda, T.; Nakamura, H.; Ishima, T.; Obokata, T.; Tsue, M.; Kono, M. P. I. Mech. Eng. D: J. Aut. 2003, 217, 935-941. (9) Zeng, K.; Huang, Z.; Liu, B.; Liu, L.; Jiang, D.; Ren, Y.; Wang, J. Appl. Therm. Eng. 2006, 26, 806-813. (10) Huang, Z. H.; Liu, L. X.; Jiang, D. M.; Ben, Y.; Liu, B.; Zeng, K.; Wang, Q. P. I. Mech. Eng. D: J. Aut. 2008, 222, 1657-1667.
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: DOI:10.1021/ef8011382
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Figure 2. Schematic diagram of the experimental setup. Table 1. Major Specification of Gasoline and CNGDI Engines item
gasoline enginea
CNGDI engine
borestroke (mm) displacement (cc) number of cylinders compression ratio combustion chamber IVO (BTDC) IVC (ABDC) EVO (BBDC) EVC (ATDC) fuel system rated power (kW@rpm) rated torque (Nm@rpm) fuel pressure (bar) valve train (cylinder configuration)
7688 1597 4 10 bowl 12° 48° 45° 10° MPI 82/6000 148/4000 3.25 DOHC 16 V(4 cylinders in-line)
7688 1597 4 14 bowl 12° 48° 45° 10° CNGDI 82/6000 148/4000 20 DOHC 16 V(4 cylinders in-line)
a
The base engine for the CNGDI engine.
equipped in accordance with SAE standard J1349 JUN95.12 Both engines were tested from 1500 to 6000 rpm with wide open throttle (WOT) condition for comparison purposes. The composition of a natural gas fuel varies with location, climate, and other factors. It is anticipated that such changes in fuel properties affect emission characteristics and performance of CNG fuel in engines as shown by refs 11 and 13. The physicochemical properties of CNG and gasoline fuels used in this experiment are shown in Tables 2 and 3, respectively. The gross calorific value and specific gravity of the used gasoline fuel (C8H18) are 47.80 MJ/kg and 0.692, respectively. The lube oil used was ordinary commercial lube oil (SAE 40). Optimization of CNGDI Engine. The CNGDI engine is optimized with consideration to three important parameters: brake thermal efficiency, brake mean effective pressure, and
The operating range of the used lambda meter is from 0.70 to 32.00 λ and the air-fuel ratio range of any gasoline engine to be measured is about 10-22. Hence, for CNGDI engine development, the used lambda meter was good enough to tune up the engine configuration to achieve maximum best torque (MBT). The used lambda meter accurately determines only one mixture strength to achieve best performance. The information for the combustion analysis system (CAS) includes control software, encoder and pressure sensors and can be seen in ref 11. Other sensors, such as a total of nine thermocouples and six pressure sensors, were installed into the engine test bed to measure temperature and pressure at various test points. The instrument used in this investigation was fully (11) Kalam, M. A. Performance and simulation study of a compressed natural gas direct injection engine; Ph.D. Thesis, University of Malaya: Malaysia, 2008. (12) SAE Handbook; Society of Automotive Engineers: Warrendale, PA, 2002; Volume 2, p 15.
(13) Byung, H. M.; Chung, J. T.; Kim, H. Y. Simsoo, P. KSME Int. J. 2002, 16 (2), 219-226.
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Table 2. Natural Gas Composition component
mol (%)
methane ethane propane isobutane normal-butane isopentane hexane carbon dioxide nitrogen others (H2O)
94.42 2.43 0.03 0.43 0.05 0.01 0.01 0.97 0.44 1.21
Table 3. Physicochemical Properties of CNG and Gasoline Fuels properties density (kg/m3) gross calorific value (MJ/kg) molecular weight (g/mol) specific gravity
CNG
gasoline
0.81 49.00
692 47.80
16.69
114.23
0.64 (compared to air)
0.692 (compared to water)
Figure 3. Brake power vs engine speed from 1500 to 6000 rpm at WOT.
NOx emissions. The air-fuel ratio, ignition, and injection timing parameters are optimized to achieve the best brake thermal efficiency, brake mean effective pressure, and lowest NOx emissions. Table A1 (in Appendix A) shows the optimization of the CNGDI engine at constant 2000 rpm with WOT. Similar optimization is done for each operating speed. It can be mentioned that the CNG fuel is injected at constant 130 BTDC (in the compression stroke), and the ignition timings are varied between 28 to 32 BTDC to maximize above three parameters. The average air-fuel ratio all over the speed range for CNG and gasoline engines are 11.64 and 13.97, respectively. The optimized air-fuel ratio for each speed can be seen in Figure A1 (Appendix A). Figure 4. Brake torque vs engine speed at WOT.
Results and Discussion
Figure 4 shows brake torque versus engine speed from 1500 to 6000 rpm for both the gasoline and CNGDI engines at WOT. It is found that gasoline and CNGDI engines produced maximum torques of 128 N m (at 4500 rpm) and 123.47 N m (at 5500 rpm), respectively. The brake torque of the CNGDI engine is 4.53 N m lower (on average all over the speed range) than that of the gasoline engine with the same displacement volume. The average brake torque over the operating speed range obtained is 120.54 and 113.36 N m by gasoline and CNGDI engines, respectively. The reason of producing lower brake torque by the CNGDI engine is mainly due to lack of chemical energy conversion to mechanical energy, which is strongly related to volumetric efficiency, fuel mixing, net heat release rate, as well as cylinder pressure. Figure 5 shows the variation (sample means of three repeated tests with their variation around the mean) of brake specific fuel consumption (BSFC) versus engine speed for both the test engines from 1500 to 6000 rpm at WOT. It can be seen that the BSFC is high at 1500 rpm for both the engines due to a high level of friction and pumping work, and the increased relative importance of friction and heat transfer, which decreases the gross indicated fuel conversion efficiency.14 It is found that the CNGDI engine shows a slightly lower SFC between 2500 and 5500 rpm as compared to the gasoline engine, which is the reason for the lower brake power within this speed range. The lowest SFC (243.34 g/(kW h)) is found for the CNGDI engine at 3500 rpm followed by the gasoline (254 g/(kW h)
The engine test room temperature was 25 °C. The CNGDI engine did not have any initial starting difficulties since a spark plug is used for ignition. In this investigation, a total of two engines have been tested, which are (1) a gasoline engine, and (2) a CNGDI engine that has been modified from a gasoline engine. Both engines have the same cylinder volume, that is, 1.6 L. The results shown in this paper are obtained from WOT with variable speed condition. All the results obtained from experimental tests are discussed as follows. Figure 3 shows brake power versus engine speed from 1500 to 6000 rpm for both the CNGDI and gasoline engines at WOT. Both engines produce their maximum brake power at 6000 rpm, which is 70.21 kW (for the gasoline engine) and 73.04 kW (for the CNGDI engine). The graph shows the sample means and variation around the mean of three repeated tests. The average brake power over the operating speed range obtained 45.37 kW (with a variation of ( 0.91%) and 47.39 kW (with a variation of (1.12%) by CNGDI and gasoline engines, respectively. The CNGDI engine produces 2.83 kW (4%) higher brake power at 6000 rpm but on average over the entire operating speed range 2.02 kW less brake power as compared to the base engine. The reason of producing lower brake power from CNGDI engine is mainly due to producing a lower brake torque, which is strongly related to the volumetric efficiency, gas inlet temperature, gas mixture distribution, AFR, as well as cylinder pressure. However, above 5000 rpm, the CNGDI engine produces higher brake power mainly due to the richer mixture and lower volumetric efficiency.
(14) Hill, P. G. J. Eng. Gas. Turbines Power 2000, 122, 141-149.
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Figure 5. BSFC vss engine speed at WOT.
Figure 7. Performance map for CNGDI engine showing contours of constant BSFC in g/(kW h).
Figure 6. Volumetric efficiency vs engine speed at WOT. Figure 8. Performance map for original gasoline engine showing contours of constant BSFC in g/(kW h).
@3500 rpm) engine. The average SFC over the entire operating speed range for CNGDI and gasoline engines are 263.26 g/ (kW h) (with a variation of ( 0.68%) and 264 g/(kW h) (with a variation of ( 0.46%), respectively. Figure 6 shows the volumetric efficiency versus engine speed at WOT. It is found that the volumetric efficiencies of CNGDI and gasoline engines are 68 and 81%, respectively. The CNGDI engine has 16% less volumetric efficiency than the original gasoline engine. Reducing the volumetric efficiency reduces engine torque and enriches the fuel-air mixture, which consequently causes incomplete combustion. The CNGDI engine reduces volumetric efficiency mainly due to the modifications such as (i) increasing compression ratio from 10 to 14, and (ii) changing port injection to direct injection. All these parameters have been changed without modification of the intake manifold, intake port, or intake valve that together restricts the flow of air into the engine cylinder. Fuel Mapping. A performance map is the common way to present the operating characteristics of an internal combustion engine over its full load and speed range to plot BSFC contours on a graph of brake torque versus engine speed. Operation of the engine coupled to a dynamometer on a test stand, over its load and speed range, generates the torque and fuel flow-rate data from which such a performance map is derived. Figures 7 and 8 show the performance map for CNGDI and gasoline engines, respectively. Contours for constant BSFC have been analyzed, and it is found that the contour structure has changed after modification of the gasoline engine to the CNGDI engine. The gasoline engine produces a minimum BSFC of 270 g/(kW h) with the speed
range of 2000-4000 rpm and load range 80-105 N m. After modification to a CNGDI system, the minimum BSFC achieved is down to 260 g/(kW h) with the speed range from 2000 to 4500 rpm and load range of 70-110 N m. With the same BSFC, such as assuming 270 g/(kW h), the CNGDI engine can be operated with wider speed and load ranges as compared to the original gasoline engine, which is beneficial. Maximum Cylinder Pressure. The cylinder pressure is measured and analyzed to observe the effect of modification from MPI to DI system. Factors affecting the peak pressure are the engine compression ratio, load, combustion duration, net heat release rate, and air/fuel ratio distribution inside the cylinder. Figure 9 shows the cylinder pressure versus cylinder crank angle at 6000 rpm and WOT. It is observed that the CNGDI engine produces a higher maximum cylinder pressure (109 bar) than the gasoline engine (91 bar). A too high or too low cylinder peak pressure is not suitable for effective pressure acting on the cylinder during the expansion stroke, and it will reduce the engine torque.15 The reasons for producing a higher cylinder peak pressure (16% higher) by the CNGDI engine are the higher compression ratio, giving rise to a higher adiabatic pressure, fuel injection in the compression stroke, and cooler inlet gas. All these factors combined reduce mixture flame speed during ignition as well as during combustion, hence the peak pressure increases. Cylinder pressures for other operating conditions are shown in the Appendix in Figures A2 and A3. (15) Heywood, J. B. Internal Combustion Engine Fundamentals, McGraw-Hill Int. ed.; C1988; pp 140-160.
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Figure 11. Oxides of nitrogen vs engine speed at WOT. Figure 9. Cylinder pressure vs crank angle at 6000 rpm and WOT.
forms due to reaction between nitrogen and oxygen at high combustion temperature, especially with high combustion flame temperature.16 The high combustion flame temperature as well as high combustion temperature is strongly associated with stoichiometric or slightly lean engine operating conditions. In this investigation, the CNGDI engine uses a suitable rich mixture to increase the power with lower NOx emission. The emission test is repeated three times, and the results are shown with the sample means for each speed as well as the variation around the sample mean. It is found that the CNGDI engine produces lower levels of NOx as the compared to gasoline engine. The level for the CNGDI engine is 809 ppm averaged over the speed ranges with a variation of (3.62%, whereas the level for the gasoline engine is 1526 ppm with a variation of (10.01%. It is very interesting that the CNGDI engine reduces the NOx emissions by 50% as compared to the gasoline engine. This is mainly due to cool gas entering into the engine cylinder, so that the overall combustion is completed at low cylinder temperature. The CNG temperature at the common rail is 16 °C, and the intake air temperature is about 35 °C, which gives a lower combustion temperature, hence the NOx is reduced. The CNGDI engine shows consistence in NOx emission with a lower variation in the repeated tests. It can be further explained that the adiabatic cylinder pressure is high due to the increase in compression ratio from 10 to 14; and the fuel is being injected in compression stroke (at a constant pressure), which lowers the adiabatic flame temperature. The combustion completes with a lower combustion flame temperature, hence the NOx is reduced. It can be summarized that cool CNG gas, higher compression ratio, and fuel direct injection in the compression stroke are the main contributions to reduce the NOx emission. Carbon Monoxide at WOT. Carbon monoxide (CO) forms during combustion with a rich fuel-air mixture when there is insufficient oxygen to fully burn all the carbon in the fuel to CO2. As CO is strongly related to rich fuel-air mixtures, SI engines are the significant sources for CO emission, because they use stoichiometric or close to stoichiometric air-fuel ratios that may divide into fuel-rich and fuel-lean zones in the cylinder during combustion. The rich zone increases the CO emission. Hence, CO emission indicates incomplete combustion of fuel. Figure 12 shows CO emission versus engine speed at WOT. It is found that the CNGDI engine produces higher CO emission above engine speeds of 3000 rpm, mainly due to increasing fuel-air ratio and lower volumetric efficiency that causes incomplete combustion. The emission test is repeated three times, and the results are shown with the sample means for
Figure 10. Unburned hydrocarbon versus engine speed at WOT.
Unburned Hydrocarbon at WOT. Unburned hydrocarbon or partially oxidized hydrocarbon emission increases if (a) the injection occurs too early, in which case the delay time increases with the result that more fuel goes to contact the relatively cool cylinder wall; or (b) injection occurs too late, in which case there may be insufficient time for completion of combustion. The latter case may be matched with the CNGDI engine as the direct injection cooled gas enters into the engine cylinder, which is the main reason for the increase in HC emission as compared to the gasoline engine. It is found (Figure 10), however, that the CNGDI engine has a higher HC emission above 3000 rpm due to a richer mixture, which is used for better brake thermal efficiency, better brake mean effective pressure, and lower NOx emissions. The stoichiometric or lean mixture produces higher NOx emissions as shown in Appendix A. The emission test is repeated three times, and the results are shown with the sample means for each speed as well as the variation around the sample mean. The CNGDI and gasoline engines produced HC emissions of 137 and 102 ppm, respectively, averaged over the speed range. The variations around the sample mean for CNGDI and gasoline engines are (2.16 and (12.10%, respectively. The reason for the lowest variation for the CNGDI engine is thought to be the DI system. The higher HC emissions by the natural gas engine matches with another investigation in ref 9. Oxides of Nitrogen at WOT. Figure11 shows NOx concentration versus engine speed at constant WOT. The NOx normally (16) Bittner, R. W.; Aboujaoude, F. W. J. Eng. Gas. Turbines Power 1992, 114, 597-601.
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3011. Furthermore, the authors thank to UPM, UKM, UiTM, UTP, and Proton Berhad, which made it possible to produce the new CNGDI engine through collaboration research works.
Appendix A Table A1. Thermal Efficiency, BMEP, and NOx Values at 2000 rpm with Optimized Conditionsay λ thermal efficiency BMEP [bar] NOx [ppm]
0.80 0.29 7.20 1215.00
0.91 0.25 7.51 373.03
1.00 0.31 7.95 1796.50
1.11 0.31 7.45 1843.80
1.20 0.31 6.69 1347.00
a Ignition and injection timings are 28 BTDC and 130 BTDC, respectively.
Figure 12. Carbon monoxide vs engine speed at WOT.
each speed as well as the variation around the sample mean. The CO emissions of CNGDI and gasoline engines are 2.17 vol % and 1.11 vol %, respectively, averaged over the speed range. The variations around the sample mean for CNGDI and gasoline engines are (3.6 and (33.3%, respectively. The CNGDI engine has a lower variation in repeated tests than the gasoline engine, which is the effect of the DI system. Conclusion The CNGDI engine did not have any initial starting difficulties since spark plugs are used for ignition. The engine did not show any combustion noise at a compression ratio of 14 (the initial compression ratio was 10). The following conclusions may be drawn from the present investigation: I The CNGDI and gasoline engines produced a maximum brake power of 73.04 and 70.21 kW, respectively, at 6000 rpm and WOT. The CNGDI produces a 4% higher brake power at 6000 rpm than the base engine. II The CNGDI engine has a 50% lower NOx emission than the original base gasoline engine. III The CNGDI engine produces higher HC and CO emissions than the base engine by 34 and 48%, respectively. IV The CNGDI engine reduces volumetric efficiency by 16% as compared to the gasoline engine. V The CNGDI engine increases 16% peak pressure as compared to the gasoline engine.
Figure A1. Air-fuel ratio vs engine speed at WOT.
Figure A2. Cylinder pressure vs crank angle at 3000 rpm and 50% throttle.
Drawback. The CNGDI engine was developed through the modification of a gasoline engine. However some components are not modified, such as the intake and exhaust manifolds that reduce volumetric efficiency and consequently require richer fuel-air mixtures to obtain higher brake power. In the next stage, we will modify intake and exhaust manifolds to improve volumetric efficiency in order to achieve higher brake power, higher efficiency, and lower emissions with lean CNG fuel combustion. A verification, validation, and lean combustion study will be conducted. Acknowledgment. The authors thank Mr. Sulaiman Bin Ariffin (Laboratory Assistant) and Muhammad Redzuan bin Umar (Research Assistant) for providing special technical assistance related to engine test bed, ECU calibration, and data collections. Without their help, it would have been very difficult to complete the engine tests. A special acknowledgement is also offered to the University of Malaya and the Ministry of Science, Technology and Innovation for research sponsor of this project through Vote- IRPA No: 33-02-03-
Figure A3. Cylinder pressure vs crank angle at 2000 rpm and 50 N m load.
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