Homogeneous charge compression ignition versus dual fuelling for utilizing biogas in compression ignition engines

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1 Homogeneous charge compression ignition versus dual fuelling for utilizing biogas in compression ignition engines S Swami Nathan, J M Mallikrajuna, and A Ramesh* Department of Mechanical Engineering, IC Engines Laboratory, Indian Institute of Technology Madras, Madras, Chennai, India The manuscript was received on 19 July 2008 and was accepted after revision for publication on 5 December DOI: / JAUTO Abstract: In this work, biogas was used in a compression ignition (CI) engine in the homogeneous charge compression ignition (HCCI) mode as well as in the dual-fuel mode together with diesel. In the HCCI mode, the charge temperature and amount of diesel injected into the intake manifold were used to control combustion. The presence of carbon dioxide in biogas suppresses the high heat release rates normally encountered in neat-diesel-fuelled HCCI engines. Efficiencies close to diesel operation together with extremely low levels of nitric oxide (NO) and smoke were attained in a brake mean effective pressure (BMEP) range from 2.5 bar to 4 bar in the biogas diesel homogeneous charge compression ignition (BDHCCI) mode. Proper control over the charge temperature was essential. Thermal efficiency was higher and NO, hydrocarbon, carbon monoxide, and smoke levels were lower than in the biogas diesel dualfuel mode. Thus, the BDHCCI mode is a viable option for using biogas in CI engines in the medium-load ranges. Operation of the engine in the CI mode with diesel below a BMEP of 2.5 bar, then in the HCCI mode up to a BMEP of 4 bar, and in the dual-fuel mode at higher BMEPs could lead to good overall performance and low emissions. Keywords: biogas, dual fuel, homogeneous charge compression ignition, engine, alternative fuels, nitrogen oxide control 1 INTRODUCTION Biogas can be produced by the anaerobic digestion of waste animal and plant matter. The plant used to produce biogas is relatively simple to build and operate. Biogas is usually generated in organic waste treatment plants in urban areas and also in smaller units in remote locations for decentralized power production. Biogas contains approximately two thirds (by volume) of methane (CH 4 ) and the rest is mostly carbon dioxide (CO 2 ) with traces of hydrogen sulphide (H 2 S). The presence of CO 2 lowers the calorific value, flame velocity, and flammability limits of biogas. CH 4 which is the main constituent has a high self-ignition temperature. Thus biogas is resistant to autoignition *Corresponding author: Department of Mechanical Engineering, IC Engines Laboratory, Indian Institute of Technology Madras, Madras, Chennai, , India. nathan_icengines@ yahoo.co.in and knock in spark ignition (SI) engines. Further, the CO 2 that is present suppresses knock. Biogas needs a relatively small amount of air for combustion per unit mass of fuel in comparison with other hydrocarbon (HC) fuels. The important properties of biogas are given in Table 1. In general, SI engines can easily use biogas. However, the thermal efficiencies are low and high HC levels are also seen because of problems arising from the CO 2 that is present [1]. Removal of CO 2 from biogas leads to an increase in the flame velocity and calorific value and thus improves thermal efficiency as well [2]. However, this is a cumbersome process and may be difficult to adopt in small systems. Additives that increase the flame velocity such as hydrogen (H 2 ) have also been used to improve SI engine performance when biogas is used [3]. However, these methods need another fuel or devices. It is not viable to use biogas in compression ignition (CI) engines directly on account of its high JAUTO970 F IMechE 2009 Proc. IMechE Vol. 223 Part D: J. Automobile Engineering

2 414 S Swami Nathan, J M Mallikrajuna, and A Ramesh Table 1 Biogas properties [1] Calorific value 17 MJ/kg Density at 1 atm and 15 uc 1.2 kg/m 3 Flame speed 0.25 m/s Stoichiometric air-to-fuel ratio 5.7 ((kg air)/kg fuel) Flammability limits vol % in air Research octane number 130 Autoignition temperature 650uC self-ignition temperature. However, it is usually used in the dual-fuel mode when a mixture of biogas and air is inducted into a diesel engine in the normal way. This charge is then compressed to normal compression ratios used in a diesel engine and then a small amount of diesel is injected near top dead centre (TDC) to initiate combustion of the biogas. The dualfuel engine can use a wide variety of gaseous fuels in this way. When biogas is used in this mode, thermal efficiencies are always lower than the neat diesel CI mode of operation. This is again due to the presence of CO 2 [4]. In the case of fuels such as liquefied petroleum gas (LPG) or H 2 the thermal efficiencies in the dual-fuel mode can be higher than the CI mode. HC levels are considerably higher in the dual-fuel mode than in the CI mode. In the case of biogas this is a significant problem [5, 6]. Although smoke levels are lowered significantly in the dual-fuel mode in comparison with the diesel-based CI mode, the carbon monoxide (CO) and HC levels increase significantly. In addition, the dual-fuel mode of operation is suitable only at high loads. This is because, at low loads, the amount of pilot diesel injected becomes too small, which creates a poor ignition source for the lean gas air mixtures that are inducted at these conditions [7]. The dual-fuel engine leads to high levels of nitrogen oxides NO x with most fuels. Although NO x emission is reduced with biogas in the dual-fuel mode owing the presence of CO 2, the levels are still considerable. On the whole, CO 2 in biogas affects the performances of SI and dual-fuel engines. 1.1 Homogeneous charge compression ignition In a homogeneous charge compression ignition (HCCI) engine, a premixed charge of air and fuel is compressed to a high compression ratio. The charge temperature or composition (through additives or diluents) is altered to achieve autoignition near TDC. Ignition occurs simultaneously at many points in the combustion chamber. Therefore, there is neither a diffusion flame nor a clear flame front [8, 9]. The charge is very lean and this leads to relatively low temperatures and hence extremely low levels of NO x [10]. On the other hand, this method of combustion also results in high HC levels in comparison with SI or conventional CI modes. Control of combustion in HCCI engines is very difficult. High charge temperatures or fuels with a low self-ignition temperature could lead to too early combustion and knock [11]. However, diluted charge and fuels with a high selfignition temperature could lead to too late combustion. Combustion control through combination of fuels has been demonstrated by several researchers. Natural gas with dimethyl ether, LPG with diethyl ether, iso-octane, and n-heptane have been used in HCCI engines for achieving good combustion control [12 14]. However, high combustion rates are always seen when the brake mean effective power (BMEP) is elevated, limiting the window of operation in the HCCI mode. CO 2 can be used as a diluent to control combustion in HCCI engines [15]. When biogas is used in the HCCI mode, CH 4 and CO 2 will prevent autoignition. Diesel can be used as an ignition improver together with higher charge temperatures. The CO 2 present in biogas can help to control combustion rate at high loads [16]. This mode of operation with biogas and diesel has not been reported in literature. Hence a detailed study will help to evaluate and compare the HCCI mode with the dual-fuel mode for the feasibility of using biogas in a CI engine. 1.2 Present work This experimental work compares the performance, emissions, and combustion characteristics of a CI engine when using biogas as the main fuel while being run in two modes, namely HCCI and dual fuel. In the duel-fuel mode, biogas was inducted and diesel was injected in the conventional way. In the HCCI mode, biogas was inducted and diesel was injected into the manifold to aid ignition. In addition, the charge temperature was also increased in the HCCI mode to achieve stable ignition. A comparison has been made between these modes of operation as regards performance, emissions, and combustion at different loads and biogas-to-diesel energy ratios. 2 EXPERIMENTAL SET-UP The base engine used in this work is a single-cylinder water-cooled direct-injection diesel engine which was modified to operate in the HCCI mode, and the schematic diagram of the set-up is shown in Fig. 1. The specifications of the engine are seen in Table 2. An electrical heater was placed in the intake air stream to control the charge temperature. There Proc. IMechE Vol. 223 Part D: J. Automobile Engineering JAUTO970 F IMechE 2009

3 HCCI versus dual fuelling for utilizing biogas in CI engines 415 Fig. 1 Schematic diagram of the experimental set-up: 1, biogas floating drum; 2, biogas fuel control and measurement system; 3, manifold injection system; 4, intake manifold; 5, 6, heater system; 7, airflow measurement; 8, pressure transducer; 9, crank angle encoder; 10, data acquisition system; 11, personal computer; 12, cooling-water inlet; 13, water flow measurement; 14, cooling-water temperature measurement; 15, cooling-water outlet; 16, exhaust gas temperature measurement; 17, exhaust manifold; 18 21, exhaust gas analysers; 22, dynamometer controller; 23, eddy current dynamometer; 24, clutch controller; 25, electric motor; 26, electromagnetic clutch; 27, engine were no changes in the combustion chamber geometry to convert it from the CI to the HCCI mode. However, a separate in-line injection system run by the cam shaft which could be set at any injection timing desired was developed and used. The required injection timings were set with the help of an external mechanical arrangement. A needle-lift sensor was used to find the actual start of injection. The pintle-type diesel fuel injector was placed after the electrical heater. Biogas generated in a fixed dome reactor was collected in a flexible bag and then transported to the laboratory. It was then sent into a floating drum meant to maintain a constant pressure. The flowrate of the biogas was controlled by a valve and was measured with a gas flow meter before being admitted into the intake manifold. Intake air and exhaust gas temperatures were measured by means of the thermocouples and the Table 2 Engine specifications Bore 80 mm Stroke 110 mm Connecting-rod length 231 mm Compression ratio 16:1 Rated output 3.7 kw at 1500 r/min Displacement volume m 3 Injection opening pressure 220 bar coolant water outlet temperature was measured using a resistance temperature detector. An optical encoder and a flush-mounted water-cooled, piezoelectric pressure transducer were used to measure the in-cylinder pressure history. During experiments, in-cylinder pressure data were acquired for 100 consecutive cycles. The heat release rate (HRR) was calculated from the cylinder gas pressure history based on a method proposed in the literature [17]. A flame ionization detector for HCs, a non-dispersive infrared analyser for CO, a chemilumiscent analyser for NO x, and a Bosch smoke meter for smoke emissions were used. 3 EXPERIMENTAL PROCEDURE 3.1 HCCI mode of operation First, the engine was motored with the intake charge temperature being increased steadily using the heater. Then, the engine was started in the HCCI mode with diesel as the fuel. Diesel was injected into the inlet manifold after the heater. An intake charge temperature of 85 uc was required for starting. Subsequently the engine was loaded to a BMEP of 2.5 bar, i.e. 50 per cent of the engine rated load. Then JAUTO970 F IMechE 2009 Proc. IMechE Vol. 223 Part D: J. Automobile Engineering

4 416 S Swami Nathan, J M Mallikrajuna, and A Ramesh biogas was slowly introduced together with the intake charge, and the amount of diesel that was injected was reduced simultaneously; the engine speed was maintained at 1500 r/min throughout. The coolant outlet temperature was maintained at 50 uc throughout. Three intake charge (air + biogas + diesel) temperatures namely 80 uc, 100 uc, and 135 uc were tried at different diesel biogas ratios at the same BMEP of 2.5 bar. The limits of operation in the HCCI mode were misfire (at high biogas flowrates) and knock (at low biogas flowrates). Subsequently the BMEP was raised from 2.5 bar in steps to 4 bar at different constant charge temperatures of 80 uc, 100 uc, and 135 uc. The engine could not be run at BMEPs below 2.5 bar in the HCCI mode as the intake temperature could not be raised beyond 135 uc owing to system limitations. Beyond a BMEP of 4 bar the combustion rate became very high. Results have been reported at different biogas energy ratios: biogas energy ratio 5 energy from biogas/(energy from biogas + energy from diesel) Dual-fuel mode of operation Experiments were carried out on the same engine at a constant speed of 1500 r/min in the dual-fuel mode with biogas being introduced and diesel being injected in the normal way. Diesel produced the ignition source. Here again different BMEPs were tried. At each BMEP the amount of biogas was progressively increased until misfire. The amount of diesel that was injected was automatically reduced by the governor to maintain the speed. The coolant outlet temperature was maintained at 50 uc and the intake charge (biogas + air, in this case) temperature was maintained at 35 uc. The injection timing for biogas dual-fuel operation has to be more advanced than diesel operation for best thermal efficiency as reported in literature [4, 5]. Hence, the injection timing was set at 28 u before TDC which is 5u crank angle more advanced than diesel operation. 4 RESULTS AND DISCUSSION 4.1 Comparison of the biogas diesel homogeneous charge compression ignition and the biogas diesel dual-fuel modes at a fixed load The brake thermal efficiency in the biogas diesel homogeneous charge compression ignition (BDHCCI) mode of operation at a BMEP of 3.0 bar is seen in Fig. 2 at three constant charge temperatures. The Fig. 2 Comparison of brake thermal efficiencies at a BMEP of 3.0 bar brake thermal efficiency in the biogas diesel dual-fuel (BDDF) mode at the same BMEP at a charge temperature of 35 uc is also depicted. In the HCCI mode, biogas was supplied until the efficiency dropped significantly and misfire occurred. When the charge temperature is increased, the engine could be operated in the HCCI mode with higher biogas energy ratios. However, the lower limit of the biogas energy ratio has to be maintained at a high level to avoid knock when the charge temperature is elevated. It may be noted that in the diesel HCCI mode even a charge temperature of 100 uc cannot be used as it will lead to severe knocking due to high heat release rates. With biogas, higher charge temperatures could be tolerated owing to the presence of CO 2 and the high self-ignition temperature of CH 4 present in it. Too high levels of biogas lowered the HRR significantly, also delayed the combustion process and thus lowered the thermal efficiency. At a BMEP of 3 bar the increase in the charge temperature was beneficial as seen in the Fig. 2. However, at higher BMEPs, lower charge temperatures (80 uc at a BMEP of 4 bar, and 100 uc at a BMEP of 3.5 bar) lead to the best thermal efficiency. At these conditions, high charge temperatures lead to too advanced combustion and high HRRs. The presence of CO 2 in biogas helps to suppress this condition as the biogas amount is increased. Table 3 indicates the charge temperature, biogas energy ratio, and thermal efficiency at the best condition at different BMEPs in the BDHCCI mode. Sample Table 3 BMEP (bar) BDHCCI operating conditions Charge temperature (uc) Energy ratio (%) Brake thermal efficiency (%) Proc. IMechE Vol. 223 Part D: J. Automobile Engineering JAUTO970 F IMechE 2009

5 HCCI versus dual fuelling for utilizing biogas in CI engines 417 Figure 2 indicates that the highest thermal efficiency (24.3 per cent) is obtained at an energy ratio of 63 per cent. In the BDDF and BDHCCI modes the thermal efficiency is lower than for neat diesel (energy ratio, 0 per cent) operation. The biogas dual-fuel mode literature also indicates that the thermal efficiency always decreases as the amount of biogas is increased [4]. This is because the CO 2 present in biogas lowers the flame velocity. However, the BDHCCI mode offers higher thermal efficiencies than the BDDF mode. Probably still higher thermal efficiencies and biogas rates could have been achieved if the charge temperature was elevated further at this BMEP. However, this was not tried in this work owing to system limitations. In the BMEP range bar, as will be shown later, the BDHCCI mode was better then the BDDF mode. BDHCCI operation beyond this BMEP range was not possible because of the high HRR. The charge temperatures needed for the BDHCCI mode can be obtained through heating by the exhaust gas. The variations in HC emission (non-ch 4 HC 5 were not measured) at a BMEP of 3.0 bar in the BDHCCI mode (three different charge temperatures) as well as in the BDDF mode are seen in Fig. 3. In the BDHCCI mode, as the amount of energy from biogas increases, the HC level increases probably because the unburned portion of the charge, which is thought to be composed of biogas, is richer. Beyond a particular substitution there is sudden rise in the HC level due to an increase in partial-burn and misfiring cycles. The HC levels are far higher than the base diesel values. BDDF engines also show significantly increased HC levels in comparison with normal diesel operation. This is a generally observed trend even with other gaseous fuels in the case of dual-fuel engines. The reason is that in the dual-fuel mode the introduced mixture is lean and flame quenching in the combustion chamber at crevices and near walls will lead to higher HC levels. Further, improper combustion caused by ignition by small pilot fuel quantities can lead to high HC levels. NO levels are too low in the BDHCCI mode of operation at all energy ratios and charge temperatures, as seen in Fig. 4. This is due to lean homogeneous combustion of biogas. However, in the BDDF mode of operation, even though there is a drop in the NO level when the biogas energy ratio is increased, the values are far higher even when the charge temperature is only 35 uc. It is seen that the NO emission decreases as the amount of biogas introduced increases owing to the presence of CO 2. Figure 5 indicates that the CO level rises as the biogas energy level is increased beyond a particular value in the BDHCCI mode. Near the operating condition of highest thermal efficiency the CO level is low. Partial combustion at high biogas flows (low Fig. 4 Comparison of NO emissions at a BMEP of 3.0 bar Fig. 3 Comparison of HC emissions at a BMEP of 3.0 bar Fig. 5 Comparison of CO emissions at a BMEP of 3.0 bar JAUTO970 F IMechE 2009 Proc. IMechE Vol. 223 Part D: J. Automobile Engineering

6 418 S Swami Nathan, J M Mallikrajuna, and A Ramesh diesel input) could lead to increased CO. In the BDDF mode the CO level also increases with increasing biogas energy level because of a rise in the equivalence ratio and also a reduction in the pilot diesel quantity. Figure 6 shows the impact of BDHCCI and BDDFoperationonsmokeemission.IntheBDDF mode, smoke is reduced drastically as the energy ratio increased owing to a reduction in the amount of diesel injected. In the BDHCCI mode, smoke levels are always extremely low and well below the diesel and dual-fuel modes of operation. This is because both the diesel and the biogas are premixed. Simultaneous reduction in NO and smoke is a significant advantage in the BDHCCI mode in comparison with the BDDF mode. In the duel-fuel mode the heterogeneous nature of the charge formed owing to the injected diesel is the reason for the higher smoke and NO levels. Figure 7 shows the HRR in the BDDF mode at biogas energy ratios in the range 0 77 per cent at a charge temperature of 35 uc and a BMEP of 3.0 bar. It is found that the addition of biogas reduces the HRR. This is because of the presence of CO 2 and also because the CH 4 present in biogas has a high selfignition temperature. Further, the start of combustion is delayed as the amount of biogas inducted increases. This is because the introduction of biogas also affects the thermodynamic properties of the compressed charge; it leads to a lower temperature during the compression process in comparison with the neat fuel operation. In addition, the introduction of biogas also lowers the concentration of oxygen in the charge and increases the ignition delay of the pilot fuel [4]. At very high biogas energy rates the engine misfires. Fig. 7 Effect of the biogas energy ratio on the HRR in the BDDF mode Figure 8 shows the HRR in the BDHCCI mode at a charge temperature of 135 uc, in the range of biogas energy ratios that were possible (59 72 per cent) at a BMEP of 3.0 bar. Here it is found that the addition of biogas delays the combustion and also reduces the HRR because of the presence of CO 2 as well as CH 4. At the lowest energy ratio where the amount of diesel is relatively high, it can be seen that there is a small heat release before the main combustion phase. This is due to the cool flame. Higher amounts of biogas do not show this effect as CH 4 is known not to exhibit this phenomenon [18]. The highest thermal efficiency occurred at a biogas energy rate of 63 per cent owing to optimal phasing of combustion. Even a 9 per cent increment in biogas energy ratio reduces the HRR peak significantly (from 70 J/deg to 40 J/deg). A crank angle very low biogas energy ratio (less than 63 per cent) leads to early combustion at a high rate, which reduces the brake thermal efficiency, as seen earlier in Fig. 2. At this condition, as the charge Fig. 6 Comparison of smoke emissions at a BMEP of 3.0 bar Fig. 8 Effect of the biogas energy ratio on the HRR in the BDHCCI mode Proc. IMechE Vol. 223 Part D: J. Automobile Engineering JAUTO970 F IMechE 2009

7 HCCI versus dual fuelling for utilizing biogas in CI engines 419 temperature is reduced the HRR decreases with a simultaneous reduction in thermal efficiency. At a BMEP of 3 bar at each charge temperature there is an optimum biogas energy ratio, as seen in Fig. 2. The HRR in these conditions is given in Fig. 9. It can be seen that all of these have similar combustion phasings with differences in the peak, which is due to the changes in the charge temperature as well as the biogas-to-diesel ratio. 4.2 Comparison of the BDHCCI and the BDDF modes at different loads Similar experiments were conducted in the BDHCCI and BDDF modes at different BMEPs. In the case of the BDHCCI mode the best conditions as regards the charge temperature and biogas energy ratio were identified. These are indicated in Table 3. The best operating condition is normally around a biogas energy level of 50 per cent. The engine could only operate between 2.5 bar and 4 bar in the BDHCCI mode. In the case of the BDDF mode the thermal efficiency reduces as the biogas ratio is increased at all BMEPs. Hence, for a comparison between the two modes of operation a biogas energy ratio of 50 per cent was chosen. Figure 10 compares the brake thermal efficiency of BDHCCI and BDDF modes of operation with the neat diesel CI mode of operation. As mentioned earlier, in the BDHCCI operation the best points (at conditions shown in Table 3) are plotted. The brake thermal efficiencies of the BDHCCI mode as well as the BDDF mode are lower than in neat diesel CI mode operation. However, the BDHCCI mode is always better than the BDDF mode owing to higher energy release rates, which can be seen in Figs 7 and Fig A comparison of the HRRs of the BDHCCI and BDDF modes is seen in Fig. 11. As seen in Fig. 12, the HC emissions (non-ch 4 HC 5 were not measured) are significantly lower in the BDHCCI mode. This is due to mode complete Fig. 11 Comparison of brake thermal efficiencies at different loads Comparison of HRRs at a BMEP of 3 bar at a biogas energy ratio of 50 per cent Fig. 9 Effect of the charge temperature on the HRR in the BDHCCI mode with optimum biogas energy ratios (CA, crank angle) Fig. 12 Comparison of HC emissions at different loads JAUTO970 F IMechE 2009 Proc. IMechE Vol. 223 Part D: J. Automobile Engineering

8 420 S Swami Nathan, J M Mallikrajuna, and A Ramesh combustion on account of the higher charge temperatures. However, it is higher than the CI mode. Higher charge temperatures could have lowered the HC levels in the BDDF mode also but would have had an adverse effect on NO emissions. At high BMEPs, i.e. 3.5 bar and 4.0 bar, the HC levels are comparable for the BDHCCI and BDDF modes. Figure 13 shows the NO emissions. Even at elevated temperatures such as 135 uc, NO emissions are too low in the BDHCCI operation but in the BDDF mode they are high even at a charge temperature of 35 uc owing to the non-uniform distribution of diesel. HCCI operation always results in less than 20 ppm NO. CO levels are also lower in the BDHCCI mode in comparison with the BDDF mode owing to complete combustion, as seen in Fig. 14. Higher biogas rates mean a lower amount of air (since biogas is introduced as a gas and displaces air) and hence the introduced charge becomes richer. This will raise the CO levels. In addition, a high biogas means lower quantity of injected diesel (on a per energy basis, the quantity of diesel which was required for 50 per cent of the total energy is very much lower). Since the injected diesel forms the ignition source in the BDDF mode, this could affect complete combustion of the biogas and lead to high HC and CO levels. Dissociation of CO 2 in biogas also could lead to higher CO levels at high BMEPs. As expected, the lowest smoke levels are seen in the BDHCCI mode because of the homogeneous nature of the charge (Fig. 15). At the BMEP of 4 bar the smoke levels are 2.4 Bosch smoke units (BSU), 0.6 BSU, and 0.1 BSU in the CI, BDDF, and BDHCCI modes respectively. Since the engine cannot be started in the biogas diesel HCCI mode, it was started in the diesel HCCI mode. Thus the manifold injected diesel HCCI mode of operation was only used to start the biogas diesel HCCI mode. The same could have also been achieved through an in-cylinder diesel HCCI mode or the duel-fuel mode. Since the engine was started in diesel HCCI mode, comparisons were made in Table 4. Manifold diesel HCCI mode suffered from poor brake thermal efficiency and high HC emissions because of manifold wall wetting, fuel accumulation, and lubricant dilution [11]. The biogas diesel HCCI efficiency is high because of proper combustion phasing. 5 CONCLUSIONS Fig. 13 Comparison of NO emissions at different loads 1. In the BDDF mode, the presence of CO 2 (in biogas) lowers the thermal efficiency when the amount of biogas is increased. However, in the case of the BDHCCI mode the CO 2 present helps to avoid high HRRs that are normally experienced in diesel HCCI operation. Too high biogas flowrates, however, lead Fig. 14 Comparison of CO emissions at different loads Fig. 15 Comparison of smoke emissions at different loads Proc. IMechE Vol. 223 Part D: J. Automobile Engineering JAUTO970 F IMechE 2009

9 HCCI versus dual fuelling for utilizing biogas in CI engines 421 Table 4 Comparison of manifold diesel HCCI with the biogas diesel HCCI BMEP (bar) Brake thermal efficiency (%) HC (ppm) NO x (ppm) Smoke (BSU) CO (vol %) Mode Manifold diesel HCCI Biogas diesel HCCI Manifold diesel HCCI Biogas diesel HCCI Manifold diesel HCCI Biogas diesel HCCI Manifold diesel HCCI Biogas diesel HCCI to misfire, and low biogas flowrates result in rapid heat release and knock. 2. Efficiencies close to diesel operation together with extremely low levels of NO and smoke were attained in a BMEP range from 2.5 bar to 4 bar in the BDHCCI mode. This can be extended to still higher BMEPs with proper control over the charge temperature and biogas energy ratio. 3. The NO level was less than 20 ppm and the smoke level was less than 0.1 BSU in the BDHCCI mode. However, in the BDDF mode the NO levels were very high even at a charge temperature of 35 uc. HC emissions were significantly higher than in neat diesel mode and also rose with increase in the biogas energy ratio. However, the values were lower in the BDHCCI mode in comparison with the BDDF mode. 4. The best energy ratio of biogas was about 50 per cent in the BDHCCI mode. A charge temperature of about uc was needed. This could be attained though heating by exhaust gases. On the whole the BDHCCI mode of operation is better than the BDDF mode with respect to performance and emissions in the range tested. The operating window is, however, narrower than for the BDDF mode. It is suggested that the engine can be operated in the CI mode with diesel below a BMEP of 2.5 bar, then in the BDHCCI mode up to a BMEP of 4 bar in the BDHCCI mode, and in the BDDF mode at higher BMEPs for good overall performance and low emissions. This will be possible with a flexible fuel injection system and efficient controls for biogas. REFERENCES 1 Porpatham, E. Experimental investigations on a biogas fuelled SI engine. PhD Thesis, Indian Institute of Technology, Madras, India, Porpatham, E., Ramesh, A., and Nagalingam, B. Investigation on the effect of concentration of methane in biogas when used as a fuel for a spark ignition engine. Fuel, 2008, 87(8 9), Porpatham, E., Ramesh, A., and Nagalingam, B. Effect of hydrogen addition on the performance of a biogas fuelled spark ignition engine. Int. J. Hydrogen Energy, 2007, 32, Bari, S. Effect of carbon dioxide on the performance of biogas/diesel dual fuel engine. Int. J. Renewable Energy, 1996, 9(3), Prakash, G., Ramesh, A., and Tazerout, M. Influence of injection timing and load on the performance and combustion characteristics of a biogas/diesel dual fuel engine. Fuels Int., 2001, 1(3), Narayana Reddy, J. Experimental investigations on a biodiesel fuelled CI engine. MS Thesis, Indian Institute of Technology, Madras, India, Poonia, M. P., Ramesh, A., and Gaur, R. R. Experimental investigation of the factors affecting the performance of a LPG dual fuel engine. SAE paper , Onishi, S., Jo, S. H., Shoda, K., Jo, P. D., and Kato, S. Active thermo-atmospheric combustion (ATAC) a new combustion process for internal combustion engines. SAE paper , Najt, P. M. and Foster, D. E. Compression-ignited homogeneous charge combustion. SAE paper , Ryan, T. W. and Callahan, T. J. Homogeneous charge compression ignition (HCCI) of diesel fuel. SAE paper , Swami Nathan, S., Mallikarjuna, J. M., and Ramesh, A. Effect of coolant and charge temperatures in a manifold injected premixed charge compression ignition engine an experimental study. In Proceedings of the International Conference on IC engines and combustion, Hyderabad, India, December 2007, paper F131 (JNTU Publishers, Hyderabad, India). 12 Shigeyuki, T., Ayala, F., James, C. K., and Heywood, J. B. Two-stage ignition in HCCI combustion and HCCI control by fuels and additives. Combust. Flame, 2003, 132, Chen, Z., Konno, M., and Goto, S. Study on homogenous premixed charge CI engine fueled with LPG. JSAE Rev., 2001, 22, Kim, D. S. and Lee, C. S. Improved emission characteristics of HCCI engine by various premixed fuels and cooled EGR. Fuel, 2006, 85, Sahashi, W., Azetsu, A., and Oikawa, C. Effects of N 2 /CO 2 addition on ignition and combustion in homogeneous charge compression ignition engine JAUTO970 F IMechE 2009 Proc. IMechE Vol. 223 Part D: J. Automobile Engineering

10 422 S Swami Nathan, J M Mallikrajuna, and A Ramesh operated on dimethyl ether. In Proceedings of the Sixth International Conference on Modeling and diagnostics for advanced engine systems (COMO- DIA 2004), Yokohama, Japan, 2 5 August 2004, pp (JSME, Yokohama, Japan). 16 Swami Nathan, S., Mallikarjuna, J. M., and Ramesh, A. Effect of adding N 2 and CO 2 on control of combustion in a manifold injected homogeneous charge compression ignition engine an experimental study. In Proceedings of the 2008 IEEE International Conference and Expo on Multimedia (ICME 2008), Johur Bahru, Malaysia, 2008, pp (IEEE, New York). 17 Brunt, M., Rai, H., and Emtage, A. L. The calculation of heat release energy from engine cylinder pressure data. SAE paper , Glassman, I. Combustion, 3rd edition, 1996 (Academic Press, London). APPENDIX Notation BDDF BDHCCI BMEP BSU CH 4 CI CO CO 2 HC HCCI HRR H 2 S LPG NO x SI biogas diesel dual fuel biogas diesel homogeneous charge compression ignition break mean effective pressure Bosch smoke units methane compression ignition carbon monoxide carbon dioxide hydrocarbon homogeneous charge compression ignition heat release rate hydrogen sulphide liquefied petroleum gas nitrogen oxide spark ignition Proc. IMechE Vol. 223 Part D: J. Automobile Engineering JAUTO970 F IMechE 2009