Effect of mixing chamber venturi, injection timing, compression ratio and EGR on the performance of dual-fuel engine operated with HOME and CNG

Abstract

Stringent environmental policies and the ever increasing demand for energy have triggered interest in novel combustion technologies that use alternative fuels as energy sources. Of these, pilot-ignited compressed natural gas (CNG) engines that employ small biodiesel pilot to ignite a premixed natural gas–air mixture have received considerable attention. This paper discusses the effect of mixing chamber venturi, injection timing, compression ratio and exhaust gas recirculation (EGR) on the performance of dual-fuel engine operated on biodiesel derived from honge oil and is called honge oil methyl ester (HOME) and CNG. The proposed study mainly focuses on the manifold induction of CNG along with HOME injection. However, CNG can also be injected using port or direct gas injector (Lakshmanan and Nagarajan 2010, Energy 35, pp. 3172–3178). The future study will involve these methods of CNG injection. From this study, it is concluded that an advanced injection timing and an increased compression ratio resulted in increased brake thermal efficiency and reduced smoke, hydrocarbons and carbon monoxide emissions. However, nitrogen oxides (NO _x ) emission increased significantly. The increased NO _x emission was effectively reduced with EGR method. A mixing chamber venturi of 3 mm size, injection timing of 27° before top dead centre (BTDC), compression ratio of 17.5 and 10% EGR were found to be optimum for the modified compression ignition engine that was operated on CNG–HOME dual-fuel mode.

Keywords:

• honge oil
• honge oil methyl ester
• emissions
• mixing chamber venturi
• compression ratio
• injection timing

1. Introduction

Petroleum resources are finite and therefore search for their alternative non-petroleum fuels for internal combustion engines is continuing all over the world (Papagiannakis and Hountalas 2003). The use of alternative gaseous fuels in diesel engines is increasing worldwide. The use of gaseous fuels is prompted by the cleaner nature of their combustion compared with conventional liquid fuels as well as by their relatively increased availability at attractive prices. The compression ignition (CI) engines have good fuel efficiency and high power output and also have the ability to use high-quality renewable fuels that can be produced efficiently from biomass, which is carbon dioxide (CO₂) neutral. Dual-fuel mode of operation employing compressed natural gas (CNG) and biodiesels of plant oils such as honge and jatropha oils is an attractive option as our country has a large agriculture base that can be a feedstock to this fuel technology and can ease the burden on the economy by curtailing the fuel imports. Natural gas (NG) has high octane number, and is therefore suitable for engines with a relatively high compression ratio (Balasubramanian et al. 1995, Daisho and Takahashi 1995, Liu and Karim 1995, Kusaka et al. 2000, Lin and Su 2003, Papagiannakis and Hountalas 2003, Srinivasan et al. 2005, Carlucci et al. 2008). The use of CNG as an alternative fuel has far-reaching environmental and economic implications. It can be used either as a sole fuel in spark ignition engine or can be dual fuelled with liquid fuels in CI engines. Dual-fuel engines have drawn considerable research attention in the area of alternate fuels. The two main advantages of this concept are as follows: no major modifications are required in the existing engine; there is a flexibility of engine to switch back to the diesel mode of operation as and when need arise. Dual-fuel combustion system utilising combined diesel and NG fuel has been proposed in recent years (Balasubramanian et al. 1995, Daisho and Takahashi 1995, Liu and Karim 1995, Kusaka et al. 2000, Lin and Su 2003, Papagiannakis and Hountalas 2003, Srinivasan et al. 2005, Carlucci et al. 2008). Due to its low cetane number, CNG in dual-fuel mode demands pilot injection of diesel fuel as the ignition source. Use of different biodiesels as injected fuel along with CNG induction in dual-fuel mode has been reported in the literature (Barsic and Humke 1981, Selim 2005, Bahman et al. 2007, Carlucci et al. 2008, Yoshimoto 2010). This study highlights the performance of dual-fuel engine with engine parameters that have been well investigated by other researchers. However, most of the researchers investigate performance of such dual-fuel engines using diesel fuel only. From the literature survey, it follows that no significant study has been done with CNG–biodiesel dual-fuel engines. Different methods of CNG utilisation in diesel engines have been reported in the literature (Heywood 1988). These methods include gas manifold induction (Papagiannakis and Hountalas 2003, Nwafor 2007, Papagiannakis et al. 2007), port injection (Selim 2005, Selim et al. 2008, Lakshmanan and Nagarajan 2010) and direct in-cylinder gas injection (Ishida et al. 2003, McTaggart-Cowan et al. 2004). The main objective of this study involves the effect of mixing chamber venturi, injection timing, compression ratio and exhaust gas circulation (EGR) on the performance evaluation of dual-fuel engine operated on manifold induction of CNG with pilot injection of honge oil methyl ester (HOME). This study involves the use of locally available biodiesel as an injected fuel that replaces fossil diesel fuel. In the literature, little study has been done on CNG–biodiesel dual-fuel engine. Further study involves port/direct CNG injection in a biodiesel-injected dual-fuel injection with an advanced combustion analysis.

The effect of injection timing on the performance and emission characteristics of diesel engine running on NG in dual-fuel mode has been reported in the literature (Nwafor 2007, Papagiannakis et al. 2007, Sayin et al. 2008, Selim et al. 2008, Sahoo et al. 2009). It is reported that the advanced injection timing results in better engine performance. The effect of compression ratio on the performance and emission characteristics of diesel engine running on NG in dual-fuel mode has been reported in the literature (Sahoo et al. 2009). It is reported that knock starts earlier when a high compression ratio is used, and an increased compression ratio generally increases the combustion noise. The main problem with CNG–diesel/biodiesel-fuelled dual-fuel engines is higher nitrogen oxides (NO _x ) emissions. EGR is an effective technique to reduce NO _x emissions in diesel engines. The effect of EGR on the performance and emission characteristics of diesel engine running on NG in dual-fuel mode has been studied by many researchers (Ishida et al. 2003, McTaggart-Cowan et al. 2004, Srinivasan et al. 2007, Pirouzpanah and Saray Khoshbakhti 2007). It is reported that an EGR level of 30% results in a relatively low NO _x emission without degrading the combustion events.

2. Characterisation of honge oil, HOME and CNG

In this study, HOME, a biodiesel derived from the locally available honge oil is used as the injected pilot fuel and CNG as the inducted fuel. The honge oil is popularly known as pongamia or karanja oil. Among the non-edible vegetable oils, the honge oil is considered as one of the better fuels for internal combustion engines. It is also used as biodiesel when its viscosity is reduced by the method of transesterification. This oil was converted into its methyl ester known as HOME by the transesterification process. NG is produced from gas wells or tied in with crude oil production. NG is primarily made up of methane (CH₄), but frequently contains trace amounts of ethane, propane, nitrogen, helium, carbon dioxide, hydrogen sulphide and water vapour. Methane is the principal component of NG. Normally, more than 90% of NG is methane. NG can be compressed, so it can be stored and used as CNG. NG is safer than gasoline in many respects, and ignition temperature of NG is higher than that of gasoline and diesel fuel. In addition, NG is lighter than air and will dissipate upward rapidly if a rupture occurs. Gasoline and diesel will pool on the ground, increasing the danger of fire. CNG is non-toxic and will not contaminate groundwater if spilled. Properties of these fuels are listed in Tables 1 and 2.

Table 1 Properties of diesel, honge oil and HOME.

Serial no.	Properties	Diesel	Honge oil	HOME
1	Viscosity @ 40°C (cst)	4.59	44.85	5.6
2	Flash point (°C)	56	270	163
3	Calorific value (kJ/kg)	45,000	35,800	36,010
4	Density (kg/m^3)	830	915	890
5	Type of oil	–	Non-edible	Non-edible

Table 2 Properties of NG (McTaggart-Cowan et al. 2004).

Properties	NG
Composition of NG	2.18% nitrogen
	92.69% methane
	3.43% ethane
	0.52% carbon dioxide
	0.71% propane
	0.12% isobutene
	0.15%, n-butane
	0.09% pentane
	0.11% hexane
Boiling range (K @ 101,325 Pa)	147
Density (kg/m^3 at 1 atm and 15°C)	0.77
Flash point (K)	124
Octane number	130
Flammability limits/range
Rich	0.5873
Lean	1.9695
Flame speed (cm/s)	33.80
Net energy content (MJ/kg)	49.5
Auto-ignition temperature, K (°C)	923 (650)
Combustion energy (KJ/m^3)	24.6
Vapourisation energy (MJ/m^3)	215–276
Stoichiometric A/F (kg of air/kg of fuel)	17

3. Experimental set-up

The engine tests were conducted on a four-stroke, single-cylinder, water-cooled, direct injection CI engine with a displacement volume of 662 cc, compression ratio of 17.5:1, developing 5.2 kW at 1500 rev/min. Figure 1 shows the overall view of the test rig modified to operate on dual-fuel mode.

Figure 1 Compression ignition engine test rig with dual-fuel arrangement.

The specifications of the engine are given in Table 3. The engine was always operated at a rated speed of 1500 rev/min. The engine had a conventional fuel injection system. The injector had three holes each of 0.27 mm in diameter.

Table 3 Specifications of the CI engine.

Serial no.	Parameters	Specifications
1	Machine supplier	Apex Innovations Pvt. Ltd, Sangli, Maharashtra, India
2	Type	TV1 (Kirloskar)
3	Software used	Engine soft
4	Nozzle opening pressure	200–225 bar
5	Governor type	Mechanical centrifugal type
6	No. of cylinders	Single cylinder
7	No. of strokes	Four stroke
8	Fuel	H.S. Diesel
9	Rated power	5.2 kW (7 hp) @ 1500 rpm
10	Cylinder diameter (bore)	0.0875 m
11	Stroke length	0.11 m
12	Compression ratio	17.5:1
Air measurement manometer
13	Make	MX 201
14	Type	U-type
15	Range	100–0–100 mm
Eddy current dynamometer
16	Model	AG10
17	Type	Eddy current
18	Maximum	7.5 kW at 1500–3000 rpm
19	Flow	Water must flow through dynamometer during the use
20	Dynamometer arm length	0.180 m
21	Fuel measuring unit (range)	0–50 ml

The amount of fuel injected in the engine cycle varied from 0.45 to 0.65 kg/h. The injector opening pressure and the static injection timing as specified by the manufacturer were 205 bar and 23° before top dead centre (BTDC), respectively. The governor of the engine was used to control the engine speed. The engine was provided with a hemispherical combustion chamber with overhead valves operated through push rods. Cooling of the engine was accomplished by circulating water through the jackets on the engine block and the cylinder head. A piezoelectric pressure transducer was mounted with the cylinder head surface to measure the cylinder pressure.

The proposed study involves the effects of mixing chamber venturi type, injection timing, compression ratio and EGR on the dual-fuel engine performance and emissions operated on CNG–HOME. Two mixing chamber venturis of 3 and 6 mm hole diameters were fabricated and used for the study. Figure 2 shows two types of mixing chamber venturis or carburettors used, which were designated as venturis 1 and 2. Mixing chamber venturis 1 and 2 are each of 6 and 3 mm hole size and were designed and fabricated in-house suitable for the study. Each mixing chamber has 12 holes. To study the effects of injection timing and compression ratio on the engine performance, suitable arrangements were provided in the test rig. EGR arrangement was suitably designed and fabricated and was fitted to the engine. Figure 3 shows the EGR arrangement.

Figure 2 Types of mixing chamber venturis with different hole sizes.

Figure 3 EGR arrangement.

The EGR rate is accurately calculated using the following equation:

Brake thermal efficiency (BTE) for dual-fuel engine operation is calculated by using the following equation:

where BP = brake power, kW; m _CNG = mass of CNG, kg/s; m _HOME = mass of HOME, kg/s; CV_CNG = calorific value of CNG, kJ/kg; and CV_HOME = calorific value of HOME, kJ/kg.

Tables 4 and 5 show the specifications of exhaust gas analyser and the smoke meter used for the emission measurements with measurement accuracies and uncertainties. The emissions of hydrocarbons (HC) and NO _x are measured in parts per million, while those of carbon monoxide (CO) and CO₂ were measured in percentage. The smoke is measured in Hartridge Smoke Unit (HSU).

Table 4 Specifications of exhaust gas analyser.

Type	DELTA 1600S
Object of measurement	CO, CO2 and HC
Range of measurement	HC = 0–20,000 ppm as C3H8 (propane)
	CO = 0–10%
	CO2 = 0–16%
	O2 = 0–21%
	NO x  = 0–5000 ppm (as nitric oxide)
Accuracy	HC = ± 30 ppm HC
	CO = ± 0.2% CO
	CO2 = ± 1% CO2
	O2 = ± 0.2% O2
	NO x  = ± 10 ppm NO
Resolution	HC = 1 ppm
	CO = 0.01% V
	O2 = 0.1% Vol.
	NO x  = 1 ppm
Warm-up time	10 min (self-controlled) at 20°C
Speed of response time	Within 15 s for 90% response
Sampling	Directly sampled from tail pipe
Power source	100–240 V AC/50 Hz
Weight (g)	800
Size (mm)	100 × 210 × 50

Table 5 Specifications of smoke meter.

Type	Hartridge Smoke Meter-4
Object of measurement	Smoke
Measuring range opacity (%)	0–100
Accuracy	± 2% relative
Resolution (%)	0.1
Smoke length (m)	0.43
Ambient temperature range (°C)	− 5 to +45
Warm-up time	10 min (self-controlled) at 20°C
Speed of response time	Within 15 s for 90% response
Sampling	Directly sampled from tail pipe
Power supply	100–240 V AC/50 Hz
	10–16 V DC @ 15 amp
Size (mm)	100 × 210 × 50

4. Results and discussions

The experimental investigations were carried out on a single-cylinder, four-stroke CI engine test rig to operate in a dual-fuel mode. The engine tests were conducted with dual fuel using CNG and HOME for both 80% and 100% load conditions. Furthermore, the tests were conducted on the dual-fuel engine operation with varying injection timings and compression ratios. The injection timing is varied from 19° to 27° BTDC, and the compression ratio is varied from 15 to 17.5. As the maximum compression ratio of the engine was limited to 17.5, it was not possible to study the engine performance beyond this compression ratio. In the next phase, two types of mixing chamber venturis with 3 and 6 mm hole geometries were used for the dual-fuel engine operation. Finally, the effect of EGR on NO _x reduction was studied by varying EGR rate from 5% to 20% in steps of 5%.

4.1 Effect of mixing chamber venturi on the performance of CNG–HOME dual-fuel operated engine

This section provides the effect of mixing chamber venturi on the performance of CNG–HOME dual-fuel operated engine. The engine is operated at a constant compression ratio of 17.5, and the injection timing was maintained at 27° BTDC. The injector nozzle opening pressure was maintained at 220 bar for the HOME.

4.1.1 Performance and emission analysis

Table 6 shows the comparative performance of dual-fuel engine when used with two different mixing chamber venturis.

Table 6 Comparative performance of dual-fuel engine fitted with two mixing chamber venturis.

	Carburettor or mixing chamber venturi 1		Carburettor or mixing chamber venturi 2
Parameter of interest	Load 80%	Load 100%	Load 80%	Load 100%
BTE (%)	22.7	23.5	24.3	25.8
Smoke opacity (HSU)	62	65	58	63
HC (ppm)	69	63	75	68
CO (%)	0.155	0.185	0.140	0.165
NO x	950	995	975	1020

From Table 6, it follows that BTE is higher with mixing chamber venturi 2 compared with mixing chamber venturi 1. This is because the mixing chamber venturi 2 provides better air–CNG mixing.

The smoke opacity is less with mixing chamber venturi 2. This is mainly due to better combustion occurring inside the engine cylinder.

The NO _x emissions are more for mixing chamber venturi 2. This is mainly because with this venturi the premixed combustion will be more, resulting in higher heat release rate. This higher emission of NO _x may be controlled by EGR.

The HC and CO emissions are found to be lesser for mixing chamber venturi 2. This is mainly because BTE is also higher with this venturi associated with lower HC emissions.

4.2 Dual-fuel mode of operation on CNG–HOME with variable injection timing

This section provides the effect of injection timing on the performance of CNG–HOME dual-fuel operated engine. The engine is operated at a constant compression ratio of 17.5 with mixing chamber venturi 2 in the inlet manifold. The injector nozzle opening pressure was maintained at 220 bar for the HOME.

4.2.1 Performance and emission analysis

4.2.1.1 Brake thermal efficiency

Figure 4 shows that for CNG–HOME fuel combinations for 80% and 100% loads, as the injection timing is advanced from 19° to 27° BTDC, BTE increases for both the loads. The reason for this increase in BTE is that more time will be available for CNG fuel burning and these results in better performance and improved BTE. The engine performance was smooth, and the ignition delay reduced through the advanced injection timing but tended to incur a slight increase in fuel consumption (Nwafor 2007). At 100% load, the BTE is lower compared with 80% engine operation. An improvement in BTE was achieved by advancing the injection timing. Maximum pressure and pressure rise rate are higher for the advanced injection timing (Sahoo et al. 2009).

Figure 4 Variation of BTE with injection timing for 80% and 100% loads.

4.2.1.2 Smoke opacity

Figure 5 shows the variation of smoke opacity with injection timing for both 80% and 100% loads, respectively. The smoke opacity decreases with an increase in injection timing. This is because of better combustion prevailing inside the engine cylinder.

Figure 5 Variation of smoke opacity pressure with injection timing for 80% and 100% loads.

4.2.1.3 NO _x emissions

Figure 6 shows the variation of NO _x emissions with injection timings for CNG–HOME fuel combinations, respectively. As the injection timing increases, the emission of NO _x increases considerably with the dual-fuel combinations at both 80% and 100% loads, respectively. The reason for increased NO _x emissions with increased injection timing could be due to better combustion prevailing inside the engine cylinder and more heat released during premixed combustion. The higher NO _x emissions for 27° BTDC are later controlled by the appropriate use of EGR method. The variations in NO _x emissions follow changes in adiabatic flame temperature. These effects also vary with injection timing, suggesting that reaction zone stoichiometry and post-combustion mixing are also influenced by fuel composition (Badr and Abd Rabbo 2002, McTaggart-Cowan et al. 2004, 2010).

Figure 6 Variation of NO _x emissions with injection timing for 80% and 100% loads.

4.2.1.4 HC emissions

Figure 7 shows the variation of HC emissions with injection timings for CNG–HOME dual-fuel combination operation. As the injection timing increases, the emission decreases considerably as shown in the figure for both 80% and 100% loads. The reason for decreased HC emissions with an increased injection timing could be due to better combustion with increased BTE and more heat released during premixed combustion. However, other researchers reported that the advanced injection timing showed low and high HC emissions at low and high loading conditions compared with the standard injection timing operation, respectively (Nwafor 2007).

Figure 7 Variation of HC emissions with injection timing for 80% and 100% loads.

4.2.1.5 CO emissions

Figure 8 shows the variation of CO emissions with injection timings for CNG–HOME dual-fuel operation for both 80% and 100% loads, respectively. As the injection timing is advanced from 19° to 27° BTDC, the CO emission also decreased considerably as shown in the figure. The emission of CO results from incomplete combustion of HC fuel. The emission of CO greatly depends on the air–fuel ratio relative to stoichiometric proportions. The reason for decreased CO emissions with increased injection timing could be due to better combustion with increased BTE. The advanced injection timing showed a significant reduction in CO emissions compared with standard dual-fuel operation (Nwafor 2007).

Figure 8 Variation of CO emissions with injection timing for 80% and 100% loads.

4.3 Dual-fuel mode of operation on CNG–HOME with variable compression ratio

This section provides the effect of compression ratio on the performance of CNG–HOME dual-fuel operated engine. The engine is operated at a constant injection timing of 27° BTDC with mixing chamber venturi 2 in the inlet manifold. The injector nozzle opening pressure was maintained at 220 bar for the HOME.

4.3.1 Performance and emission analysis

4.3.1.1 Brake thermal efficiency

Figure 9 shows the variation of BTE for CNG–HOME dual-fuel operated engine at 80% and 100% loads, respectively. It is observed that BTEs increase with an increase in compression ratio from 15 to 17.5, the reason for this increase in BTE is due to better combustion taking place inside the engine cylinder. The maximum BTE was observed for a compression ratio of 17.5, and its value is found to be 24.23% in case of 80% load. The BTEs observed for CNG–HOME dual-fuel mode of operation at 15, 16 and 17.5 compression ratios were 17%, 18% and 24.23%, respectively. Increase in the compression ratio generally increases the combustion noise due to the higher self-ignition possibility of the gaseous fuels at higher pressures and temperatures. As the compression ratio was reduced, the combustion noise also reduced with an extended ignition limit (Selim et al. 2008, Sahoo et al. 2009).

Figure 9 Variation of BTE with compression ratio for 80% and 100% loads.

4.3.1.2 Smoke opacity

Figure 10 shows the variation of smoke with CNG–HOME dual-fuel engine operation at 80% and 100% loads, respectively, for different compression ratio. The smoke opacity reduces with an increase in compression ratios. The main reason for this decrease in smoke is mainly due to better combustion prevailing inside the engine cylinder. Similar observations were recorded even for full-load engine operations. It is a well-known fact that duel-fuel operation remarkably reduces smoke emission. Moreover, combustion of CNG produces no particulates, the only smoke associated is mainly due to pilot injection of diesel/biodiesel (Mahla et al. 2010).

Figure 10 Variation of smoke opacity with compression ratio for 80% and 100% loads.

4.3.1.3 NO _x emissions

Figure 11 shows the variation of NO _x emissions with CNG–HOME dual-fuel operated engine at 80% and 100% loads, respectively. It is observed that NO _x emissions increase with an increase in compression ratios. The reason for increased NO _x emissions with an increased compression ratio is because of more heat released during premixed combustion. The NO _x emissions are controlled by proper EGR% for 17.5 compression ratio. The formation of NO _x is favoured by high oxygen concentration in the biodiesels used and high charge temperature (McTaggart-Cowan et al. 2004, Bahman et al. 2007). Higher NO _x emissions for injected vegetable oils in dual-fuel operation may be due to the availability of higher oxygen that is present in the molecular structure of the injected vegetable oils the molecular structure. Also for the same power output more vegetable oils are required (due to increased specific fuel consumption), leading to delayed injection. For CNG–HOME combinations, the exhaust gas temperature increases. This high peak temperature increases NO _x emissions (McTaggart-Cowan et al. 2004, Bahman et al. 2007).

Figure 11 Variation of NO _x with compression ratio for 80% and 100% loads.

4.3.1.4 CO emissions

Figure 12 shows the variation of CO emissions with CNG–HOME dual-fuel operated engine at 80% and 100% loads, respectively. It is observed that CO emissions decrease with an increase in compression ratio. The reason for decreased CO emissions with an increased compression ratio is because of higher BTE observed at these loads. The rate of CO formation is a function of the unburned gaseous fuel availability and mixture temperature, both of which control the rate of fuel decomposition and oxidation (Papagiannakis and Hountalas 2003, Bahman et al. 2007, Mahla et al. 2010).

Figure 12 Variation of CO with compression ratio at 80% and 100% loads.

4.3.1.5 HC emissions

Figure 13 shows the variation of HC emissions with CNG–HOME dual-fuel operated engine at 80% and 100% loads, respectively. It is observed that HC emissions decrease with an increase in compression ratio. The reason for decreased HC emissions with an increased compression ratio is because of higher BTE. Lower HC emissions at a compression ratio of 17.5 are observed. The variation of unburned hydrocarbons in the exhaust gases is consistent with the quality of the combustion process of the engine (Bahman et al. 2007, Nwafor 2007, Selim et al. 2008). With an increase in engine load, there is a sharp decrease in HC emissions under dual-fuel operation. This is the result of the increase in burned gas temperature that helps to oxidise efficiently the unburned hydrocarbons (Papagiannakis and Hountalas 2003, McTaggart-Cowan et al. 2004, Papagiannakis et al. 2007, Sahoo et al. 2009, McTaggart-Cowan et al. 2010).

Figure 13 Variation of HC with compression ratio for 80% and 100% loads.

4.4 Effect of EGR on the performance of CNG–HOME dual-fuel operated engine

This section provides the effect of EGR on the performance of CNG–HOME dual-fuel operated engine. The engine is operated at a constant injection timing of 27° BTDC and a constant compression ratio of 17.5 with mixing chamber venturi 2 in the inlet manifold. The injector nozzle opening pressure was maintained at 220 bar for the HOME. The EGR is varied from 5% to 20% in steps of 5%. EGR beyond 20%, although lowers NO _x emissions, significantly dilutes the charge reducing the BTE with increased smoke, HC and CO emissions.

4.4.1 Performance and emission analysis

4.4.1.1 Brake thermal efficiency

From Figure 14, it is observed that with the EGR levels of 5% and 10%, there is a small improvement in BTE. This could be due to an improved combustion. Beyond 15% EGR level BTE reduces significantly. This could be due to predominant dilution effect of EGR leaving more exhaust gases in the combustion chamber. As the EGR amount is increased from 0% to 15%, increase in BTE is observed at all loads (80% and full loads). There are three effects of using EGR in a diesel engine, namely dilution effect, chemical effect and thermal effect (Selim 2005, Srinivasan et al. 2007). This could be due to improved an combustion of CNG. The inlet temperature increases when the EGR is introduced beyond 15% EGR. The chemical effect is associated with the participation of active free radicals present in exhaust gas to enhance combustion by taking part in pre-ignition reactions. However, this effect causes an increase in thermal efficiency. With more EGR substitution, the thermal efficiency falls. This is due to the dilution effect of the EGR used, as it depletes the oxygen present in the combustion chamber (Mahla et al. 2010).

Figure 14 Variation of BTE with EGR for 80% and 100% loads.

4.4.1.2 Smoke

Figure 15 shows that the smoke opacity is higher in case of 20% EGR and is least in 0% EGR. Use of EGR has a negative effect on smoke emissions. The main reason for this is the reduction of engine air/fuel ratio supplied to the engine. The main advantage of a dual-fuel engine is the reduction of smoke emission. It is revealed that dual-fuel operation using NG is a very efficient method to reduce soot concentration at almost all conditions. The main reason is that NG, for which methane is the main constituent, being a lower member in the paraffin family, has very small tendency to produce soot. Thus, practically, gaseous fuel produces no soot, although it contributes to the oxidation of the soot formed from the combustion of the liquid fuel. Thus, the use of NG has a positive effect on soot emission reduction (Papagiannakis and Hountalas 2003, Sahoo et al. 2009). Recirculation of exhaust gases into combustion chamber of engine causes minor increase in smoke density of exhaust gases particularly at higher loads. During part-load operation, there are no noticeable changes in smoke density of engine.

Figure 15 Variation of smoke with EGR for 80% and 100% loads.

4.4.1.3 NO _x emissions

Figure 16 shows a reduction in NO _x emission with an increase in EGR percentage. This is because EGR reduces the oxygen concentration in the charge and reduces the temperature of combustion products due to higher specific heat capacity and hence lower NO _x is observed (Ladommatos et al. 1998, Papagiannakis and Hountalas 2003, Sahoo et al. 2009, Mahla et al. 2010). In dual-fuel operation, NO _x emission is generally low but with higher substitution of CNG higher heat release rates occur, resulting in higher emission of NO _x . Also, the oxygenate of HOME fuel combination in CNG results in higher NO _x emissions due to more complete combustion than the ester alone. All the dual-fuel combinations show increased NO _x emissions with an increase in the gas flow rate. This is mainly due to the increased quantity of heat output and the maximum temperature prevailing inside the engine cylinder with an increased gas flow rate. Thus, the combustion process is complete, adequately resulting in increased NO _x emission.

Figure 16 Variation of NO _x with EGR for 80% and 100% loads.

4.4.1.4 HC emissions

Figure 17 shows the variation of HC emissions with EGR. As EGR increases, both the oxygen concentration in the charge and the temperature of combustion products decrease. The increase in HC emission is observed with an increase in EGR percentage. This is due to poor combustion resulting inside the combustion chamber. Normally, dual-fuel operation exhibits higher emission of unburned hydrocarbon at light loads. At light loads, the pilot quantity being small flame cannot propagate fast and is far enough to ignite the entire mixture. As a result, it causes higher HC emissions, but with an increase in load the hydrocarbon emission decreases. As load progresses, the pilot quantity increases and burns the surrounding fuel–air mixture sufficiently (Ishida et al. 2003, Srinivasan et al. 2007, Sahoo et al. 2009, Mahla et al. 2010).

Figure 17 Variation of HC emissions with EGR for 80% and 100% loads.

4.4.1.5 CO emissions

Figure 18 shows that carbon monoxide emissions increase with an increase in exhaust gas regulation. Higher values of CO were observed beyond 15% EGR. The reduction in the oxygen concentration is the main cause of CO and HC emissions.

Figure 18 Variation of CO emissions with EGR for 80% and 100% loads.

The dual-fuel operation yields higher CO emissions at light load conditions. At low loads, most of the fuel left unburnt leads to poor combustion and ignition results in higher CO emissions. At higher loads, the CO emission is lower than the normal diesel operation because of better utilisation of fuel. This is mainly due to high gas temperature and faster combustion rates. With EGR substitutions, the CO formation is higher at full load condition than the dual-fuel operation without EGR. At full load condition, the availability of oxygen required for the complete combustion decreases with an increase in EGR ratio. Hence, the CO emission is higher with EGR addition in dual-fuel operation at higher load than at the light load condition (Ishida et al. 2003, McTaggart-Cowan et al. 2004, Srinivasan et al. 2007, Sahoo et al. 2009, Mahla et al. 2010). The lower volatility and higher viscosity of biodiesel affect the atomisation process, resulting in rich mixture formation. Due to quenching of the flame that travels into the rich mixture, an increased HC is resulted. The lower concentration of oxygen also enhances CO levels at fuel-rich pockets.

4.5 Combustion analysis

In this section, combustion analysis for the effect injection timing and EGR on the performance of CNG–HOME-fuelled dual-fuel engine has been presented.

4.5.1 Effect of injection timing on CNG–HOME dual-fuel engine

Figure 19 shows the cylinder gas pressure plotted for three injection timings of 19°, 23° and 27° BTDC. The cylinder pressure shows a slight decrease in peak pressure value with the retarding of fuel injection timings. Moreover, it can be seen that crank angle corresponding to the peak cylinder pressure becomes retarded at low cylinder pressure. The slow combustion rate is considered to be the main reason for low cylinder pressure (Singh et al. 2004).

Figure 19 In-cylinder pressure versus crank angle for different injection timings at 100% load.

Figure 20 shows the heat release rate with different fuel injection timings. Similar to the behaviour of cylinder pressure versus injection timings, heat release rate decreases with the retarding of injection timing. The injection timings retarded yielded lower efficiencies (Singh et al. 2004). The most retarded injection timings resulted in combustion delayed into the expansion stroke. The phasing of the rate of heat release curves away from top dead centre combined with combustion occurring at low temperatures resulted in a loss in fuel conversion efficiency (Singh et al. 2004). The most advanced injection timings with their very long ignition delay period result in more fuel prepared for burning at the start of combustion compared with the retarded injection timings. Simultaneous ignition of a large proportion of the HOME fuel leads to fast combustion and a short burning time, as shown in Figure 19.

Figure 20 Rate of heat release versus crank angle for different injection timings for 100% load.

4.5.2 Effect of EGR on CNG–HOME dual-fuel engine

Figure 21 shows the cylinder gas pressure plotted for EGR rates of 5%, 10%, 15% and 20%. The cylinder pressure shows a slight decrease in peak pressure value with an increase in EGR rate. The slow combustion rate is considered to be the main reason for low cylinder pressure.

Figure 21 In-cylinder pressure versus crank angle for different EGR ratios at 100% load.

Figure 22 shows the heat release rate with EGR rates of 5%, 10%, 15% and 20%. Similar to the behaviour of cylinder pressure versus EGR%, the heat release rate decreases with an increase in EGR%. Moderate EGR fractions resulted in a slight increase in ignition delay. At higher EGR fractions, the combustion was significantly delayed, with the peak heat release rate occurring much later. With injection timing held constant, the later combustion indicates that the ignition delay has been substantially increased (McTaggart-Cowan et al. 2004).

Figure 22 Rate of heat release versus crank angle for different EGR ratios at 100% load.

5. Conclusions

Based on the exhaustive experimentation on CNG–HOME dual-fuel engine operation with optimised engine parameters, better efficiency and acceptable emission levels have been determined. The following conclusions were made with reference to each of the above engine parameters that were considered for the study.

Effect of gas–air mixing chamber suggests that the 3 mm size venturi results in better performance in terms of higher BTE and lower emissions.

Effect of injection timing suggests that with the advancement from 19° to 27° BTDC, the BTE increased and smoke opacity, HC and CO emissions decreased. On the other hand, NO_x emission increased and is found to be maximum at 27° BTDC. NO_x can be controlled using the EGR technique.

Effect of compression ratio shows that with the increasing compression ratio from 15 to 17.5, the BTE increased for both 80% and 100% loads. The smoke opacity, HC and CO emissions decreased. On the other hand, the NO_x emissions increased. The optimum compression ratios are found to be 17.5.

Effect of EGR shows that the BTE increased up to 10% EGR and reduced beyond 15% EGR. This is due to deterioration of combustion inside the engine cylinder.

Finally, it can be concluded that 3 mm size mixing chamber venturi, injection timing of 27° BTDC, compression ratio of 17.5 and 10% EGR are found to be optimum for the modified CI engine operated on CNG–HOME dual-fuel mode.