Effect of injection timing and EGR on engine-out-responses of a common-rail diesel engine fueled with neat biodiesel

Abstract

Nowadays, diesel-powered engines are becoming attractive worldwide due to their superior fuel economy, higher efficiency and excellent reliability. Biodiesel can be considered as the most promising and in demand alternative fuel because it is a non-toxic, biodegradable and renewable fuel. This work attempts to simultaneously reduce the BSNO_x and smoke from the levels of fossil diesel by using palm methyl ester (PME) biodiesel. In addition, this paper describes the conversion of a common-rail injection system with a custom-made electronic control system, focusing on hardware development, the engine control unit and fuel delivery system development. Parametric studies dealing with injection timing and exhaust gas recirculation (EGR) variation using neat palm biodiesel were performed and compared with baseline diesel. The tests were performed at a constant speed and load of 1500 rpm and 0.4 MPa, respectively. Firstly, the start of injection (SOI) timing was varied from TDC to −25° ATDC to demonstrate the flexible control of the custom-made engine controller. Later, the SOI timing was kept at an optimum of −11° ATDC and the EGR rates were adjusted (i.e. 0–50%). The experimental results indicated that both the injection timing and EGR variation had a prominent effect on the engine performance, emissions and combustion characteristics for an engine operating with baseline diesel or neat biodiesel. Based on the highest brake thermal efficiency (BTE) and a reasonable NO_x level, the optimum injection timing is found to be at −11° ATDC for both the baseline diesel and biodiesel operation. A wider range of EGR rates from 0% to 50% were investigated to bring down the NO_x levels from the EURO II limit to meet with more stringent EURO limits. It was found that with the PME fuel, engine operation at 30% EGR resulted in the optimum trade-off between BSNO_x and smoke emissions. In fact, simultaneous BSNO_x and smoke reduction from the levels of fossil diesel is possible with the use of PME biodiesels in parallel with the implementation of late SOI timing or a higher EGR rate in diesel engines.

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1. Introduction

Environmental protection, energy efficiency and conservation have become important issues in recent decades because of the rapid depletion of fossil fuel reserves, the rising price of crude oil, and environmental degradation. This energy crisis has triggered a revolution in the use of alternative fuels that are environmentally acceptable, economically viable, domestically available and technically feasible for internal combustion (IC) engine applications. Among the various alternative fuels, biodiesel is a prominent fuel that has immense potential to meet the world’s energy demands and facilitate the reduction of harmful pollutant emissions. Biodiesel is a renewable, nontoxic, biodegradable and oxygenated fuel. It can be produced from straight vegetable oils, edible and non-edible plants, recycled waste cooking oils, and animal fat through the transesterification process.^1,2 Furthermore, biodiesel can be flexibly used in compression ignition (CI) engines without modification of the engine or fueling process, thus greatly simplifying the system integration and adoption of its use.^3,4 However, biodiesel poses some variations in its physiochemical properties than fossil diesel fuel, such as cetane number, heating value, viscosity, density, cloud points, pour points, etc. The combined effects of these properties strongly affect injection characteristics, air–fuel mixing and the atomization of biodiesel fuel in diesel engines. Insufficient in-cylinder air motion, poor atomization and low volatility of biodiesel lead to the lower thermal efficiency, poorer combustion and higher emissions of biodiesel-fueled diesel engines. In general, engines fueled with biodiesel have reduced emissions such as carbon monoxide (CO), smoke, unburned hydrocarbons (HC), volatile organic compounds (VOCs), and the overall life cycle emissions of carbon dioxide (CO₂). However, despite the potential merits of reducing these pollutants, numerous studies revealed that the NO_x emissions will be increased with the use of biodiesel.^5–8

1.1. Strategies to improve diesel engine pollutant emissions

In CI engines, the majority of studies on the comparisons between diesel and biodiesel have been based on the standard setting of an engine using fossil diesel fuel. Also, it is generally agreed upon that the formulation of fuel composition can enhance the biodiesel combustion performance and tailpipe emissions. However, the experimental results indicated that it was not easy to keep NO_x emissions neutral while reducing other pollutants simply through fuel reformulation.^9–12 Therefore, modification of the engine operating parameters such as injection strategies and exhaust gas recirculation (EGR) may be possible to optimize the engine emissions due to the difference in combustion characteristics and chemical composition between diesel and biodiesel. In a diesel engine, improvements in the fuel injection parameters can be employed to reduce engine emissions and improve the fuel economy. Injection parameters such as injection pressure, injection duration, injection timing, and fueling are the key injection parameters which can significantly affect the performance and emissions of an engine. For instance, the combustion efficiency and ignition delay will change as the injection timing is varied because of the effect of mixture formation.¹³ Numerous studies revealed that injection timing retardation reduces NO_x emissions.^14–16 With late injection timing, the peak cylinder pressure decreases and results in lower peak combustion temperatures and consequently, NO_x emissions diminish. Conversely, advancing the injection timing decreases HC and CO emissions. In another study,¹⁷ the effect of injection timing on the performance, combustion and emission parameters was investigated in a single cylinder, mechanical pump-line-nozzle injection system using algal oil methyl ester (AOME) blended fuels (i.e. 5, 10, & 20% blend). The test result revealed that the advancement in injection timing of 5 CAD from the rated static injection timing of 345 CAD caused a reduction in brake specific fuel consumption, HC, CO and smoke, and increase the combustion pressure, heat release rate, brake mean effective pressure (BMEP) and NO_x emissions. In another related study, Ganapathy et al.¹⁸ have demonstrated an improvement in engine performance and emission when the fuel injection pressure and injection timing were optimized for Jatropha biodiesel operation. The experiment was conducted in a single-cylinder diesel engine that was equipped with a mechanical pump-nozzle injection system. The fuel injection timing was varied with a 5 degree crank angle on either side of the rated static injection timing (345 CAD). Another effective approach to reduce the NO_x from a petrol–diesel engine is by means of the EGR technique which is a pretreatment approach. However, using EGR alone has some drawbacks in that it could reduce energy efficiency, operational stability and a trade-off in terms of soot emissions.¹⁹ In this regard, others have investigated the effects of combining biodiesel and EGR. The general outcome from these studies was that combining EGR and biodiesel was an effective strategy to reduce NO_x and/or PM.¹² Pradeep and Sharma²⁰ adjusted the EGR levels (5–25%) and engine load on a single-cylinder engine and found that biodiesel emitted more smoke at lower loads and less smoke at higher loads when compared to diesel fuel. Tsolakis et al.²¹ found that the use of biodiesel fuel could reduce the smoke and NO_x from a single-cylinder engine equipped with EGR (i.e. 10% and 20%) under certain engine conditions when compared to diesel. FOME (fish oil methyl ester) and its blends have been tested in a diesel engine by Bhaskar et al.²² They show that a blend fuel with 20% vol of FOME produces nearly the same brake thermal efficiency with lower unburned hydrocarbon, carbon monoxide and soot emissions, but higher NO_x emissions compared to diesel fuel. They found that NO_x emissions can be reduced with the use of EGR. EGR flow-rates of 10%, 20% and 30% were examined in their study. The authors suggested that a 20% EGR flow rate is optimum for a 20% FOME blend considering the emissions of NO_x and soot.

1.2. Diesel engine modifications

The conventional problems of the mechanical type pump-line-nozzle injection system with fixed injection timing is well-known, with their low combustion efficiency and high exhaust emissions. The conventional injection system leaves little room for engine performance optimization to be tailored to biodiesel fuels since the injection process, controlled by the camshaft, is dependent on the engine speed. In a CI diesel engine, the fuel injection system (including the injection nozzles and pump) plays a vital role because it directly affects the performance of the engine. Several desired demands are: higher injection pressure, optimized injection rate, higher precision of injection timing control, and higher precision of injection quantity control, which could significantly affect the mixture formation and combustion quality of the engine. In fact, all of these injection system parameters must be controllable especially for alternative fuel research studies. Generally, commercially available single-cylinder engines are equipped with a mechanical fuel injection system and most of the injection parameters cannot be readily changed. Utilizing electronically controlled fuel injection through a common-rail injection system instead of the conventional mechanical injection permits the continuous control of injection timing and injection quantity to a high level of precision. This technology also offers the highest levels of flexibility for the control of both the injection timing and injection amount, while still yielding significantly better results than any conventional injection system. For engines equipped with common-rail injection technology, the fuel injectors are typically ECU controlled. There are various studies focused on the conversion of mechanically controlled fuel injection systems to an electronic common-rail system and efforts are being made for the design of the engine controller unit. In the study of Ergenç et al.,²³ the test engine which was initially equipped with mechanical injection has been modified and converted into a test engine with a common-rail injection system. All injectors (LPG and diesel) in their study were also controlled by programmable logic controllers (PLC) which served as an ECU for control applications. Likewise, in another study performed by Goldwine,²⁴ a dedicated common-rail injection single-cylinder air-cooled diesel engine was converted from a mechanical injection design, and most of the parts of the injection system were adopted from regular diesel engine parts. A piezoelectric type injector was employed in their study and multiple data acquisition/control cards were used as an ECU for the control of fuel injection, injection pressure and engine load. The software was written in Labview and the algorithm implemented closed loop control for the engine speed (through load regulation) and fuel pressure. The piezoelectric injector used in their study has the ability to implement up to six injections per cycle with various lengths and dwell times.

1.3. Objectives of the paper

As mentioned above, most of the previous studies into biodiesel fuel have been performed with conventional mechanical pump-line-nozzle fuel injection systems. The compressibility effects due to the changes in the physical properties of biodiesel could potentially affect the fuel injection timing and this can increase NO_x. Using a common-rail injection system could eliminate this issue because the fuel injection pressurization is not dependent on the injection timing.²⁵ Aside from the challenge of fuel injection technology, other considerations like the cost-effectiveness of biodiesel production is the new topic for debate. From an economic point of view, the use of the most cost-effective biodiesel feedstock will pave the way for the large-scale production of biodiesel. By far, the three most common available biodiesel feedstocks are palm oil (from Malaysia), soybean (from the US) and rapeseed (from the EU) and their production cost are USD $ 684, $751 and $ 996 a tonne, respectively.²⁶ Apparently, the palm oil offers the most cost competitiveness and it is a viable biodiesel feedstock. Furthermore, palm oil has been cited as a high-yield source of biodiesel with an average yield of about 5950 litres per hectare, which is nearly 13 times better than the yield of soybean oil.²⁷ Considering the cost and the promising yield of the feedstocks for biodiesel production, biodiesel derived from palm oil is clearly the most viable substitute for petroleum diesel fuel. Thus far, most of the research about biodiesel, including the study of its effect on engine performance, emissions and combustion characteristics, has been performed under relatively low EGR levels (i.e. <30%) and with an engine equipped with a conventional pump-line-nozzle injection system. The engine-out-responses under higher EGR (>30%) conditions have not yet been sufficiently investigated. In the case of palm biodiesel fuel, it contains about 11.7% oxygen content in the fuel composition and has a higher cetane number than petro-diesel,²⁸ which gives great opportunities to optimize the engine performance and emissions under higher EGR levels. Consequently, there is a strong motivation to investigate the impact of neat palm oil methyl ester (PME) combustion in a diesel engine equipped with a common-rail injection system with higher EGR rate.

Another motivation for this study is to develop a fully controlled common-rail fuel injection single-cylinder diesel engine for a wider range of renewable fuel research studies. As discussed above, this system is able to provide flexible control of injection parameters such as injection timing, injection pressure, and number of injections in a cycle of operation, which enables a more advanced combustion study. Usually, the common-rail injection system can be found in multi-cylinder diesel engines used in passenger cars and trucks. However, they are too large and complicated, and it is almost impossible to have full access to the stock ECU to reconfigure the injection parameters. Besides, it is rare indeed that a commercially available single-cylinder diesel engine is equipped with an electronically controlled fuel injection system largely due to the high cost of implementation. Considering that the major advantages of the common-rail injection system are the improvement in thermal efficiency, fuel economy, and cleaner exhaust emissions compared to a conventional mechanical system, its introduction in a single-cylinder diesel engine should be an interesting idea. Therefore, a preliminary study on injection timing optimization was carried out for both of the baseline diesel and PME fuels on a modified single-cylinder engine test rig equipped with a high-pressure common-rail injection system.

2. Experimental apparatus and procedure

2.1. Biodiesel production process

There are numerous ways to convert vegetable oil into biodiesel fuel, such as pyrolysis, microemulsion, dilution, and transesterification. Of these different conversion methods, the transesterification process is the most popular way and has been extensively used to reduce the viscosity of crude vegetable oil and convert triglycerides into esters and glycerol. Fig. 1 shows the transesterification reaction of triglycerides. A catalyst is typically employed to enhance the reaction rate and yield. As the reaction is reversible, excess alcohol is used to shift the equilibrium toward the product side (right side).

Fig. 1 Transesterification of triglycerides with alcohol.

In the present study, crude palm oil was transferred into a preheated reactor at a temperature of 60 °C. The oil was reacted with 25% (v/v oil) methanol and 1% by weight of alkali catalyst (KOH). The reaction mixture was maintained at 60 °C for 2 hours with stirring at the constant speed of 800 rpm. After the completion of the reaction, the produced methyl esters were poured into a separation funnel for 24 hours to separate the glycerol from the biodiesel. The lower layer, which consists of impurities and glycerin, was drawn off. Then, the methyl ester was washed with warm distilled water and evaporated with a rotary evaporator at 65 °C for 30 minutes to remove residual methanol and water. Lastly, the methyl ester was dried using Na₂SO₄ and filtered using qualitative filter paper to collect the final product.

2.2. Biodiesel property test

Upon the completion of the transesterification process, the fuel properties of the produced methyl ester were comprehensively examined and compared with the biodiesel standards. Table 1 contains a description of the key physicochemical properties of the converted neat PME in comparison with ASTM and EN standards. The important properties of the petroleum diesel are also listed in this table. It can be observed that the physicochemical properties of the produced biodiesel were measured and benchmarked against the biodiesel standards based on ASTM D6751 and EN14214. It appears that all of the physicochemical properties of PME are sufficient to meet the ASTM and EN biodiesel standards. In particular, the kinematic viscosity of the transesterified palm oil was substantially improved, but it was slightly higher than that of petroleum diesel. In addition, the flash point for PME was relatively higher than that of petroleum diesel and was suitable for use as a transportation fuel. However, the calorific value of the PME was lower than that of petroleum diesel. Another key property that significantly influences the engine performance, emissions, and combustion characteristics is the cetane number of fuel. It can be observed that PME has a higher cetane index than petroleum diesel fuels. Distillation characteristics also have important effects on engine combustion and performance. Typically, the distillation temperature is used as a quality check for fuel and the distribution range provides an insight into the volatility, flash point and fatty acid composition. Biodiesel tends to shift the distillation curve towards higher boiling points than petrol–diesel, especially in the T₅₀ region.²⁹ In this study, the full ranges of the distillation temperatures of the fuel samples T_x, in which “x” stands for distillation temperatures corresponding to x vol% of the distilled and condensed liquid fuel, were measured by a distillation temperature analyzer (Anton Paar ADU 5, Anton Paar Strasse 10, 8054 Graz, Austria). The cetane index (CI) of the PME and diesel fuel was calculated from the density (D) and distillation temperature T₅₀ using the following formula:³⁰

CI = 454.74 − 1641.416D + 774.74D² − 0.554(T₅₀) + 97.803(log T₅₀)² (1)

where D = density at 15 °C and T₅₀ = mid-boiling temperature, °C.

Table 1 The fuel properties of petroleum diesel and PME biodiesel

Properties	Unit	Diesel fuel	Biodiesel
			PME	Limit (ASTM D6751)	Test method
Kinematic viscosity @ 40 °C	mm^2 s^−1	3.34	4.4	1.9–6.0	ASTM D445
Density @ 15 °C	kg m^−3	851.9	876.9	880	ASTM D127
Acid number	mg KOH g^−1	0.12	0.06	0.5 max	ASTM D664
Calorific value	MJ kg^−1	45.31	39.98	—	ASTM D240
Flash point	°C	71.5	165.5	130 min	ASTM D93
Pour point	°C	1	9	Not specified	ASTM D970
Cloud point	°C	8	10	Not specified	ASTM D2500
Cold filter plugging point	°C	−2	10	Not specified	ASTM D6371
Oxidation stability @ 100 °C	hours	>50	5	3 min	EN14112
Cetane index	—	51	58	—	D976
Distillation	°C
	IBP	182.7	297.8	Distillation temperature, 90% recovered (T90) = 360 °C max	ASTM D86
	T5	212.4	325.2
	T10	225.5	326.5
	T15	236.6	326.5
	T20	246.1	326.9
	T25	254.2	327.5
	T30	261.0	327.9
	T35	268.1	328.5
	T40	274.6	328.9
	T45	280.4	329.2
	T50	286.9	329.6
	T55	293.0	330.2
	T60	299.4	330.7
	T65	305.9	331.3
	T70	312.7	332.0
	T75	320.3	333.0
	T80	327.8	334.2
	T85	336.1	335.9
	T90	346.1	339.0
	T95	361.3	343.1
	FBP	366.0	344.3

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As can be seen, the distillation temperatures of T₅₀ for PME and diesel fuel are 329.6 °C and 286.9 °C, respectively. A higher distillation temperature may shorten the ignition delay of the fuel, thus increasing the cetane number and decreasing the probability of the occurrence of knocking in diesel engines.³¹ In addition, the fatty acid composition of PME was measured by a gas chromatography/flame ionization detector (GC/FID). The GC/FID operating conditions are given in Table 2. The analysis of fatty acids was based on AOAC 996.06 official methods. The results of the fatty acid composition of PME fuel in comparison with another study are shown in Table 3. It was found that PME contained a moderate level of saturated (44.87%) and unsaturated (55.14%) fatty acids, in which the level of saturated fatty acids is almost equal to that of the unsaturated fatty acids. In fact, the distribution of fatty acid compositions is in very close agreement with the other study.³²

Table 2 The GC/FID operating conditions

Property		Specification
Carrier gas		Hydrogen
Flow rate of carrier gas		1 ml min^−1
Column		Agilent HP-88 (60 m × 0.25 mm ID, 0.2 mm)
Inlet temperature		250 °C
Initial temperature		120 °C
Initial holding time		1 minute
Oven ramp conditions	1^st ramp	10 °C min^−1 to 175 °C (hold 10 min)
	2^nd ramp	5 °C min^−1 to 210 °C (hold 5 min)
	3^rd ramp	5 °C min^−1 to 230 °C (hold 5 min)
Type of detector		FID
Split ratio		50 : 1
FID detector temperature		260 °C
Injection volume		1 μL

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Table 3 The fatty acid composition of neat PME fuel

Property	Formula	PME	PME^32
Carbon chain length distribution (wt%)
Saturated fatty acid
C4:0 (butyric acid)	C4H8O2	0.15	0
C6:0 (caproic acid)	C6H12O2	0.08	0
C8:0 (caprylic acid)	C8H16O2	0.21	0
C10:0 (capric acid)	C10H20O2	0.18	0
C12:0 (lauric acid)	C12H24O2	1.56	0.2
C14:0 (myristic acid)	C14H28O2	1.4	0.9
C15:0 (pentadecanoic acid)	C15H30O2	0.05	0
C16:0 (palmitic acid)	C16H32O2	36.74	43.7
C17:0 (heptadecanoic acid)	C17H34O2	0.1	0
C18:0 (stearic acid)	C18H36O2	4.23	4.5
C20:0 (arachidic acid)	C20H40O2	0	0.3
C21:0 (heneicosanoic acid)	C21H42O2	0.07	0
C24:0 (lignoceric acid)	C24H48O2	0.1	0
Unsaturated fatty acid
C16:1n7 (palmitoleic acid)	C16H30O2	0.19	0
C18:1n9t (elaidic acid)	C18H34O2	0.7	0
C18:1n9c (oleic acid)	C18H34O2	41.90	39.7
C18:2n6c (linoleic acid)	C18H32O2	10.03	10.0
C18:2n6t (linolelaidic acid)	C18H32O2	0.31	0
C18:3n6 (γ-linoleic acid)	C18H30O2	0.42	0
C18:3n3 (linolenic acid)	C18H30O2	0.19	0
C20:1 (cis-11-eicosenoic acid)	C20H38O2	0.19	0
C20:2 (cis-11,14-eicosadienoic acid)	C20H36O2	1.13	0
C20:3n6 (cis-8,11,14-eicosatrienoic acid)	C20H34O2	0.08	0
Fatty acid saturation/unsaturation ratio (wt%/wt%)		44.87/55.14	49.6/49.7

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2.3. Engine operating conditions

In this study, all experiments were conducted under a constant speed of 1500 rpm and an injection pressure of 600 bar. Generally, the test program in the present experiment comprises two series of tests to assess the effects of biodiesel fuel on the engine performance, emissions, and combustion characteristics. Firstly, the effect of the fuel injection timing on performance, emissions and combustion characteristics of an engine operating in conventional compression ignition mode was investigated. At a constant BMEP of 0.4 MPa and without EGR, the start-of-injection (SOI) timing was varied from 0° ATDC to −25° ATDC. In the later test series, the EGR rate was varied from 0% to 50% at a constant BMEP of 0.4 MPa and an SOI of −11° ATDC. This SOI timing was confirmed based on the peak brake thermal efficiency as found in the previous test series. In each series of tests, diesel fuel was used as the baseline fuel for the basis of comparison. When the engine was fueled with biodiesel fuel, the engine ran satisfactorily throughout the entire test, which was performed at room temperature, and had no starting difficulties. The tests were performed under steady-state conditions with a sufficiently warmed exhaust gas and water coolant temperature. To enhance the accuracy of the study, each test point was repeated twice to produce average readings. The repeatability was matched over 95% for each test.

2.4. Electronic fuel injection system

2.4.1. Hardware and fuel delivery system development. The test engine used in the present study is based on a modified single-cylinder compression ignition diesel engine (YANMAR TF-120E). Originally, the engine was equipped with a regular fuel injection system consisting of a mechanical type of pressure fuel pump (200 bar) timed by the camshaft, and a mechanical fuel injector with an injection angle of 150° and four 0.26 mm diameter holes. The engine is naturally aspirated with a maximum output of 7.5 kW. The original fuel injection timing was constant and set to 17° BTDC. The engine was originally equipped with a mechanical governor to control the engine speed. The specifications of the test engine are listed in Table 4 and a schematic diagram of the experimental setup is shown in Fig. 2. Fig. 3 is a photo showing the common-rail injector mounted on the engine head. The original injection system of the engine was disassembled and a new common-rail injection system was retrofitted as illustrated in Fig. 4. The system was based on commercially available common-rail diesel engine components. A second-generation, electronically controlled common-rail high-pressure injection system (DENSO HP3) was installed to replace the original mechanical type pump-line-nozzle injection system. The fuel pump was externally driven using an electric motor running at 750 rpm to maintain the required high-pressure levels in the fuel rail and to ensure a stable line pressure with minimum fluctuation. The pressure level in the rail is fully controlled by PID closed-loop regulation and it can be changed to a pressure between 0 and 1800 bar.

Table 4 Characteristics of single-cylinder engine

Parameter	Before modification	After modification
Model	YANMAR TF-120E
Displacement	638 cm^3
Bore	92 mm
Stroke	96 mm
Compression ratio	17.7 : 1
Rated power	7.8 kW
Rated speed	2400 rpm
Fuel injection system	Mechanical cam driven injection	Electronically common-rail injection
Nozzle holes	4	5
Nozzle diameter	0.26 mm	0.134 mm
Fuel pump mounting	Engine mechanically driven	Electrically driven

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Fig. 2 Schematic diagram of the experiment setup.

Fig. 3 Common-rail injector mounted on the engine cylinder head.

Fig. 4 Common-rail fuel delivery setup with pump driven by electric motor.

2.4.2. Engine controller unit (ECU). The ECU employed in the present study was based on the open-source Arduino Mega 2560 microcontroller.³³ The microcontroller uses three interrupt service routines to pick up the incremental encoder and engine camshaft signals. The programming coding was written with the open-source Arduino Software (IDE) and loaded to the board via serial communication with the computer. For the real-time control and monitoring of the injection parameters such as engine speed, start of injection (SOI) timing and opening pulse-width (PW) for the pilot, main and post injection, closed-loop engine speed control mode selection and injection pressure adjustment a LabVIEW based graphic user interface (GUI) program was employed in this study. The interface programme is shown in Fig. 5. The engine controller also featured a programmable peak and hold pulse-width-modulation (PWM) to effectively drive the solenoid injectors for common-rail direct injection. The control unit was designed to fully support and control the engine parameters. The same controller system was capable of simultaneously controlling the exhaust gas recirculation (EGR) system. EGR was adopted to moderate the heat release rate (HRR) and the combustion timing phasing. In particular, this involved the installation of the EGR valve, EGR cooler, EGR surge tank and two identical CO₂ sensors. The EGR rate can be flexibly adjusted by controlling the EGR valve. Under steady-state conditions, the EGR rate can be measured by comparing the ratio of the CO₂ in the intake to the exhaust and as follows:

Fig. 5 Interface programme.

2.5. Instrumentation

The engine load absorber is based on the 7.5 kW A.C. synchronous dynamometer. It is used to provide loading to the engine and to maintain the engine speed. An airflow meter turbine with a 2 to 70 litres per second (L s⁻¹) measuring range was installed to measure the intake airflow rate. To monitor the exhaust gas temperature, a type K thermocouple was used and mounted in the exhaust stream. The fuel flow rate for the direct injection system was measured with a positive displacement gear wheel flow meter, which interfaced with a flow rate totalizer. The test system was installed with the necessary sensors for combustion analysis and fuel injection timing identification. In-cylinder gas pressure was measured using a Kistler 6125B type pressure sensor. The charge signal output of the pressure sensor was converted to a low-impedance voltage signal using a PCB model 422E53 in-line charge converter; this unit was powered using a PCB model 480E09 signal conditioner. To acquire the top dead centre (TDC) position and crank angle signal for every engine rotation, an incremental quadrature rotary shaft angle encoder with 0.125° CA resolution (X4 encoding) was used. To determine and verify the SOI timing and injection duration for the injector, the injector current signal was measured with a hall effect current sensor. To simultaneously acquire the cylinder pressure signal, injector current signal, and encoder signals, a computer equipped with a high-speed simultaneous sampling data acquisition card, which has 14 bits resolution, a 2 MS s⁻¹ sampling rate, and four analog input channels, was used. The acquired data were further processed and analysed with Matlab software. In each test, 100 consecutive combustion cycles of pressure data were collected and an average was calculated. To reduce noise effects, smooth data using SPAN as the number of points used to compute each element was applied to the sampled cylinder pressure data. Combustion parameters, such as peak pressure magnitude, peak pressure location, heat release rate, peak heat release rate location, and ID, were all computed using Matlab software. For the exhaust emission measurement, an AVL DICOM 4000 gas analyser was used to measure the concentrations of NO_x. The opacity of smoke was measured using an AVL DiSmoke 4000. All emissions were measured during steady-state engine operation. The measurement range and resolution of both of the instruments are provided in Table 5. The NO_x emissions were converted into brake specific emissions by using the following equations according to SAE J177:

BSNO_{x(g kW⁻¹ h⁻¹)} = 0.0952 × NO_x(ppm) × exhaust mass flow rate_{(kg min⁻¹)}/brake power_(kW) (3)

Table 5 Measuring components, ranges and resolution of the AVL DICOM 4000 gas analyzer and DiSmoke 4000 smoke analyzer

Equipment	Measurement principle	Component	Measurement range	Resolution
Gas analyzer	Electrochemical	Nitrogen oxides (NOx)	0–5000 ppm	1 ppm
	Calculation	Excess air ratio (λ)	0–9999	0.001
Smoke opacimeter	Photodiode detector	Opacity (%)	0–100%	0.10%

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3. Calculation methods

3.1. Engine performance

The engine performance in this work was evaluated based on the BSFC and BTE. The BSFC and BTE were determined and calculated according to the following equations:

3.2. Combustion analysis

HRR analysis is a useful approach to assess the effects of the fuel injection system, fuel type, engine design changes, and engine operating conditions on the combustion process and engine performance.³⁴ Given the plot of HRR versus crank angle, it is easy to identify the start of combustion (SOC) timing, the fraction of fuel burned in the premixed mode, and differences in the combustion rates of fuels.³⁵ In the present paper, different fuels were used in an identical compression ignition engine; hence, the HRR information is an important parameter in interpreting the engine performance and exhaust emissions. In this study, the averaged in-cylinder pressure data of 100 successive cycles, acquired with a 0.125° crank angle resolution, were used to compute the HRR. The HRR, given by , at each crank angle was obtained from the first law of thermodynamics, and it can be calculated by the following formula:where, γ = specific heat ratio, P = instantaneous cylinder pressure (Pa), and V = instantaneous cylinder volume (m³).

3.3. Statistical and equipment uncertainty analysis

Experimental errors and uncertainties can arise from instrument selection, condition, calibration, environment, observation, reading, and test procedures. The measurement range, accuracy, and percentage uncertainties associated with the instruments used in this experiment are listed in Table 6. Uncertainty analysis is necessary to verify the accuracy of the experiments. Percentage uncertainties of various parameters, such as brake specific fuel consumption (BSFC), brake thermal efficiency (BTE), and brake specific nitrogen oxide (BSNO_x) were determined using the percentage uncertainties of various instruments employed in the experiment. To compute the overall percentage uncertainty due to the combined effect of the uncertainties of various variables, the principle of the propagation of errors is considered and can be estimated as ±3.7%. The overall experimental uncertainty was computed as follows:

Overall experimental uncertainty = square root of [(uncertainty of fuel flow rate)² + (uncertainty of BSFC)² + (uncertainty of BTE)² + (uncertainty of BSNO_x)² + (uncertainty of exhaust gas temperature (EGT))² + (uncertainty of smoke)² + (uncertainty of pressure sensor)² + (uncertainty of crank angle encoder)²]

= square root of [(2)² + (1.95)² + (1.74)² + (0.73)² + (0.15)² + (1)² + (1)² + (0.03)²]

= ±3.7%

Table 6 List of measurement accuracy and percentage uncertainties

Measurement	Measurement range	Accuracy	Measurement techniques	% uncertainty
Load	±120 Nm	±0.1 Nm	Strain gauge type load cell	±1
Speed	60–10 000 rpm	±1 rpm	Magnetic pick up type	±0.1
Time	—	±0.1 s	—	±0.2
Fuel flow measurement	0.5–36 L h^−1	±0.01 L h^−1	Positive displacement gear wheel flow meter	±2
Air flow measurement	2–70 L s^−1	±0.04 L s^−1	Turbine flow meter	±0.5
NOx	0–5000 ppm	±1 ppm	Electrochemical	±1.3
Smoke	0–100%	±0.1%	Photodiode detector	±1
EGT sensor	0–1200 °C	±0.3 °C	Type K thermocouple	±0.15
Pressure sensor	0–25 000 kPa	±12.5 kPa	Piezoelectric crystal type	±1
Crank angle encoder	0–12 000 rpm	±0.125°	Incremental optical encoder	±0.03
Computed
BSFC	—	±7.8 g kW^−1 h^−1	—	±1.95
BTE	—	±0.5%	—	±1.74
BSNOx	—	±0.1 g kW^−1 h^−1	—	±0.73

---

4. Results and discussion

4.1. Effect of injection timing

4.1.1. Performance analysis. Fig. 6 illustrates the resulting effect of SOI timing on the BSFC of the engine fueled with PME biodiesel fuel and baseline diesel. Generally, the BSFC is a measure of the amount of fuel required to generate one-kilowatt of power per hour. From the results, it was observed that the BSFC for PME biodiesel fuel is consistently higher than that of baseline diesel across all SOI timings. The higher BSFC of PME means that a greater amount of fuel was required to attain the same amount of power. This was expected because of the low calorific value of PME in comparison with diesel, which was about 12% lower than that of baseline diesel fuel. Additionally, one can observe that the variation in injection timing also has a significant effect on the variation of the BSFC. As the SOI timing was advanced from the top dead center (TDC) point, the BSFC dropped for all fuels. This reduction in the BSFC can be explained by the fact that as the SOI timing was advanced, there was continuous enhancement in the combustion efficiency and quality. With a constant amount of brake power output, the decreased effect of the BSFC means less fuel is being supplied to undergo a more efficient combustion process. This is particularly for the case of advanced SOI timing. However, it is observed that a further advance in SOI timing beyond −11° ATDC causes penalties in the BSFC as the combustion pressure build up begins to resist the upward movement of the piston. Based on the BSFC results, the optimum SOI timing for the PME and diesel operations is found to be −11° ATDC and this setting will be used for the following test series.

Fig. 6 BSFC with the PME compared with diesel fuel at various SOI timing.

Engine BTE is commonly used to express the efficiency of an engine to convert fuel chemical energy to mechanical energy. BTE can be calculated by dividing the brake power output by the total energy input delivered to the system. Fig. 7 illustrates the variations in BTE with different SOI timings of the engine fueled with PME biodiesel fuel and baseline diesel. The BTE of baseline diesel is found to be consistently higher than that of PME across all SOI timing. In fact, it can be observed that the peak BTE for baseline diesel and PME are 29.5% and 28.6% respectively, at an SOI timing of −11° ATDC. In addition, the results also indicate that the BTE is significantly affected by the variation in SOI timing. There is an improvement in the BTE for all the test fuels with advanced SOI timings, except for in the case of SOI timings beyond −11° ATDC. The incremental effect is due to the longer ignition delay (physical delay) leading to better mixing, which results in better combustion and a higher BTE. Another reason is that at advanced injection timing, the engine reaches the peak pressure closer to TDC and is therefore able to produce higher effective pressure to perform useful work.³⁶ However, there is a continuous deterioration of the BTE in the case of further advances in SOI timing beyond −11° ATDC for both fuels operations. It may be due to the decrease in the delay period, which reduces the power output because a larger amount of fuel burns during expansion and the cylinder pressure rises only when the cylinder volume is expanding rapidly, and as a result lower effective pressure is produced.^36,37

Fig. 7 Brake thermal efficiency under different SOI timing conditions.

4.1.2. Emissions analysis. NO_x is a hazardous and undesirable emission product that has a wide variety of human health and environmental impacts. Literature studies indicate that there is no absolute trend in NO_x emissions when biodiesel fuels are operated in CI engines. Researchers from all over the world have reported higher NO_x emissions with biodiesel-fueled engines,^38–40 and others found lower NO_x emissions when using methyl ester fuels.^41,42 Typically, the NO_x formation depends on the fuel properties, fuel type, type of engine and engine operation conditions.^43,44 The variation of BSNO_x emissions of the test fuels at various SOI timings is illustrated in Fig. 8. The result shows that advancement of the SOI timing resulted in increased BSNO_x emissions for all the test fuels. The increasing trend in BSNO_x emissions suggested that with advanced SOI timing, the mixture ignites and burns earlier, hence resulting in early occurrence of peak pressure near TDC. This leads to a higher combustion temperature and promotes the thermal or Zeldovich NO_x formation mechanism. The results also show that PME fuel tends to lower the BSNO_x emissions across all the SOI timings. This can be attributed to the relatively higher cetane number and lower heating value of the PME compared with baseline diesel, which consequently lowers the heat release rate at the premix combustion stage and reduces the peak combustion temperature. This finding is further reinforced by the similar trend of the in-cylinder mean gas temperature, as shown in Fig. 9.

Fig. 8 BSNO_x emissions under different SOI timing conditions.

Fig. 9 In-cylinder mean gas temperature curves for (a) baseline diesel and (b) PME at various SOI timings.

The smoke formation results from the incomplete combustion of the hydrocarbon fuel and partial reaction of the carbon content in the liquid fuel. The variation of smoke emissions of the test fuels at various SOI timings is presented in Fig. 10. Generally, it can be seen that the smoke emission level decreased with the PME at all SOI timings. Lower smoke emissions were observed than in diesel fuel across all SOI timings, largely because of higher fuel-borne oxygen, lower carbon content, and the absence or lower amount of aromatics in PME fuel.⁴⁵ The results also indicate that smoke emissions were reduced with advanced SOI timings. This is due to cylinder operating temperatures being higher for advanced SOI timings, which improved the reaction between fuel and oxygen and resulted in lower smoke emissions.³⁷ Another reason may be due to the availability of sufficient time for the fuel to evaporate and mix with the air, leading to better mixing and combustion.¹⁸

Fig. 10 Smoke emissions under different SOI timing conditions.

Due to the overall lean operation and higher expansion ratio of the compression ignition diesel engine, the exhaust gas temperature (EGT) is typically lower than that of the gasoline engine. A higher EGT is unfavorable as this will deteriorate the engine fuel economy by discharging some of the useful energy into waste exhaust thermal energy, and may also cause thermal damage to piston components. The variation of exhaust gas temperature of the test fuels at various SOI timings is shown in Fig. 11. Generally, it can be seen that the variation in exhaust gas temperature follows a similar trend to the BSFC with advanced SOI timing. In fact, running on PME fuel exhibits a higher EGT when compared to diesel fuel across all SOI timing. The increment may be due to the lower calorific value of PME fuel. Thus, the increased fuel quantity injected for attaining the same amount of power has caused an increase of the in-cylinder bulk-gas-averaged temperatures. On average, EGT for PME fuel across all SOI timings were increased by 11.5 °C compared to baseline diesel. The highest increment of EGT is 15 °C for SOI of −5° ATDC with respect to baseline diesel. Another interesting observation is that as the SOI timing was advanced, EGT reduced for all fuels. This happened due to greater heat release occurring closer to TDC in the expansion stroke, which offered sufficient time for the hot combustion product to expand and cool down prior to the exhaust valve being opened. This enhances the heat utilization and allows better cooling of combustion gases, thus lowering the exhaust gas temperature. Further advances in SOI timing to beyond −13° ATDC has caused the increases in EGT because of the increase in the BSFC.

Fig. 11 Exhaust gas temperature variation under different SOI timing conditions.

4.1.3. Combustion analysis. To study the effect of biodiesel fuel on combustion, the cylinder pressures for 100 consecutive combustion cycles were recorded, averaged, and compared. Fig. 12 shows the plot of combustion pressure, HRR and injector current profile of the engine operated with baseline diesel and PME fuels at optimum SOI timing of −11° ATDC. As can be observed, the engine operated with PME fuel had little effect on the combustion characteristics, and the pattern is comparable with the baseline diesel. The pressure peak was shifted later toward the expansion stroke with PME fuel although the location of SOC timing for PME occurred 0.25° CA earlier than those of baseline diesel. In addition, a small reduction in the pressure peak in the range of 0.6 bar was observed for the PME fuel operation. Two prominent peaks of HRR were observed for both fuels. The first and second peak of HRR correspond to the premixed and mixing controlled combustion phases respectively. Also, it can be clearly observed that the location of occurrence for the first and second peak of HRR for the PME fuel was shifted earlier (by 0.25° CA) and later (by 0.875° CA) from the TDC point, respectively, compared to baseline diesel. The primary reason for the early occurrence of the first HRR peak can be attributed to the advance in SOC timing, which caused the earlier rise of the HRR. On the other hand, slow burning rate and thus longer combustion duration of PME fuel has caused the second HRR peak to occur later in the expansion stroke as compared to those of baseline diesel. The total burning angles for PME fuel with respect to baseline diesel are shown in Fig. 13. The total burning angle in this study is defined as the period between 10% and 90% mass burnt. The longer combustion duration of PME fuel means that it has a slower burn rate than baseline diesel, especially during the mixing controlled combustion phase. This was postulated to be due to the slightly higher viscosity of PME fuel compared to baseline diesel, hence delaying the mixing time required for diffusive burning. Another explanation may be due to the lower calorific value of the PME fuel, thus resulting in an increase of fuel quantity injected for attaining the same amount of power. As more fuel is being injected, a richer mixture is thus formed inside the cylinder chamber, which burns more rapidly in the early stages of combustion (premixed combustion phase) and the remaining fuel burns in the later stages (mixing controlled combustion phase) and requires a longer duration.

Fig. 12 Combustion pressure, heat release rate and injector current profiles for diesel and PME fuel at SOI of −11° ATDC.

Fig. 13 Total burning angle as a function of SOI timing for diesel and PME fuel.

Fig. 14 shows the variation of combustion pressure and HRR with respect to the crank angle at different SOI timings for the engine operated with baseline diesel and PME fuel. Generally, the combustion pressure peak consistently increases and shifted earlier toward the TDC position with advancing SOI timing for both of the fuels. The resultant higher and more effective pressure was utilized to perform useful work and thus improve the BSFC and BTE. However, further advance in SOI timing beyond −11° ATDC caused combustion pressure to build up rapidly in the compression stroke, thus beginning to oppose the upward movement of the piston and causing deterioration of the BSFC. On average, it is found that PME produces a 0.77 bar lower maximum combustion pressure compared to baseline diesel at all SOI timings. The HRR curves have similar patterns, as the combustion pressure trend where the HRR peak that is associated with premixed combustion was shifted earlier toward the compression stroke with advanced SOI timing, for both of the tested fuels. When SOI was advanced toward the TDC in the expansion stroke, the maximum HRR associated with the premixed combustion became initially lower and remained unchanged. However, further advances in SOI timing beyond −15° ATDC have led to significant increases in the maximum HRR. This was due to a longer ignition delay, which tends to promote more premixed combustion and increase both the maximum combustion pressure and HRR. With the PME fuel, the HRR was similar to baseline diesel, however one can notice that a higher fuel fraction was burned in the mixing controlled combustion phase (i.e. a wider plateau region after the first HRR peak). This phenomenon was clearly visible for the retarded SOI cases (i.e. SOI = 0° ATDC) as compared to earlier SOI conditions. This was mainly due to a higher cetane number of PME fuel compared to baseline diesel, thereby resulting in a shorter ignition delay and a lower pressure peak. In fact, the peak of the pressure curve also shifted away from the TDC point in the expansion stroke as compared to baseline diesel for the corresponding SOI timing. On average, it is found that PME produces 5.7 J/°CA lower in maximum HRR compared to baseline diesel across all SOI timing.

Fig. 14 Combustion pressure curves for (a) baseline diesel and (b) PME at various SOI timings.

4.2. Effect of EGR

EGR is one of the most promising strategies to reduce NO_x emissions in diesel engines by controlling the oxygen density and combustion peak temperature.^46,47 However, the trade-off between NO_x and soot emissions must be analyzed carefully with the addition of EGR and the biodiesel fueled engine. In this section, the effect of using EGR on the performance, emissions and combustion of the baseline diesel and PME fueled engines will be discussed. The SOI timing was kept at an optimum of −11° ATDC as found in the previous test series and adjustments were made in the EGR rates (i.e. 0 to 50%).

4.2.1. Performance analysis. The variations in the BSFC with respect to the EGR rate for the engine operated with PME biodiesel fuel and baseline diesel is shown in Fig. 15. Generally, it can be seen that the BSFC for PME biodiesel fuel is consistently higher than that of baseline diesel across all EGR rates. This was mainly due to the lower calorific value of PME, thus the BSFC is higher than that of baseline diesel at all EGR rates. One can also observe that the variation in the EGR rate also has a small effect on the BSFC. As the EGR rate increased, the BSFC dropped for all fuels, compared to without EGR. However, at an EGR rate higher than 30% and 35% for baseline diesel and PME fuel, respectively, the BSFC begins to increase gradually. At higher EGR rates, the oxygen available for combustion is reduced. Thus, the air–fuel ratio is altered and this raises the BSFC. This is evident by the decrease in excess oxygen available in the exhaust tailpipe as shown in Fig. 16.

Fig. 15 BSFC with the PME compared with diesel fuel at various EGR rates.

Fig. 16 Exhaust gas O₂ concentration with the PME compared with diesel fuel at various EGR rates.

Fig. 17 shows the comparison of the BTE for the engine operated with PME biodiesel and baseline diesel. It is evident that the BTE for PME fuel is always lower than that of baseline diesel regardless of the EGR rate. The lower calorific value of the PME fuel could be the reason behind this. Subsequently, the BTE is found to have slightly increased with a moderate EGR rate for both of the tested fuels. At moderate EGR rates, the burned gas temperature is decreased significantly, thus reducing heat loss via the combustion chamber surfaces, leaving more available for conversion to mechanical work during the expansion stroke. Another possible improvement reason may be due to the reduced pumping work as the EGR rate is increased at a constant brake load. On the other hand, lower oxygen exhaust gas feeds into the intake at higher EGR rates, thus resulting in poor air utilization and this leads to a reduction of BTE. Also, the decrease in BTE for the PME at an EGR rate of more than 35% was less prominent compared to baseline diesel. This can be credited to the higher oxygen content in PME fuel which aids the better combustion efficiency.

Fig. 17 BTE with the PME compared with diesel fuel at various EGR rates.

4.2.2. Emissions analysis. Fig. 18 shows the variation of BSNO_x emission and smoke of the PME and baseline diesel with various EGR rates. The overall trend indicates that the BSNO_x emissions for both of the tested fuels tend to decrease as the EGR rate increases. The BSNO_x emission is reduced with an increasing EGR rate due to the lowered burned gas temperature with dilution. Compared to baseline diesel, the BSNO_x emissions of PME are lower across all EGR rates. It is observed that for both of the fuels, a drastic BSNO_x reduction in the range of 23.8–97% at a 10–50% EGR rate was obtained compared with the corresponding engine operation without EGR. In fact, on average the addition of EGR in reducing BSNO_x emissions for PME was about 0.5% more effective than baseline diesel. As discussed above, a lower heat release rate during the premixed combustion phase and a lower peak combustion temperature for PME results in lower BSNO_x emissions. In addition to this, re-entering more water vapor and CO₂ into the combustion chamber due to the increase in the specific fuel consumption of PME fuel compared with the operation of baseline diesel also may lead to a greater BSNO_x reduction. This is evident by the increase in the CO₂ concentration in intake air as shown in Fig. 19. The plots also indicate that PME fuel emitted a higher exhaust CO₂ than baseline diesel across all EGR rates. This may be due to the combined effects of lower calorific value and the extra oxygen content of PME fuel altering the combustion process which eventually results in higher exhaust CO₂. At over 45% EGR, the BSNO_x emission is below 0.4 g kW⁻¹ h⁻¹ for both fuels, which is the EURO VI emission standard. On the other hand, the effects of EGR on the smoke emission for baseline diesel and PME revealed an increasing trend with higher EGR rates. It is evident that as compared to baseline diesel, the smoke emissions are lower for PME fuel and tend to increase at a much slower rate with higher EGR rates. It is the oxygen content in the PME fuel that plays a vital part in the combustion process which eventually causes a reduction in the smoke emissions. Moreover, the smoke formation rate increased sharply as the EGR rate rose over 35% for both fuels. When the engine was fueled with PME, an increase of the smoke emissions of 186% was observed using 35% EGR. Further increases in the EGR rate to 50% rapidly increased the smoke emission by 620% when compared with the engine operation without EGR. Under high EGR conditions, the exhaust gases re-circulated into the intake result in a reduction of the oxygen available for combustion. The in-cylinder soot formation and oxidation processes are strongly governed by the engine operation on gradually richer mixtures due to the reduction in oxygen content by EGR. Hence, the reduction in oxygen availability for fuel combustion and lower combustion temperature reduces the soot oxidation process which leads to higher smoke emissions. Another interesting topic that can be further discussed is the trade-off between BSNO_x, smoke and the EGR rates. From the results of the EGR effect on BSNO_x and the smoke emissions of the PME, an optimal trade-off between BSNO_x and smoke emissions can be achieved with EGR in the range of 10–30%, without a significant adverse effect on engine performance. It was realized that with PME fuel, an engine operating at 30% EGR resulted in an optimal trade-off between BSNO_x and smoke emissions. At this EGR rate, the BSNO_x emissions have effectively decreased by 80.7%, but the smoke emissions have increased by 167.3% compared to the engine operation in the absence of EGR. However, compared to diesel operation at 30% EGR, the PME fuel effectively reduced smoke emissions by 50%. Therefore, by considering both the positive effect of the reduction in BSNO_x and the smoke emissions, it is acceptable to operate an engine using PME at 30% EGR.

Fig. 18 BSNO_x and smoke emission with various rates of EGR.

Fig. 19 Intake and exhaust air CO₂ concentration with the PME compared with diesel fuel at various EGR rates.

4.2.3. Combustion analysis. To evaluate the effect of the EGR variation using PME and baseline diesel on the combustion characteristics, the in-cylinder combustion pressures for 100 consecutive combustion cycles were recorded and compared at various EGR rates (0–50%) and at a fixed engine speed of 1500 rpm and a BMEP of 0.4 MPa. The in-cylinder pressure, HRR and injector current profiles for the engine using PME and baseline diesel are both illustrated in Fig. 20. As can be seen, the variation of EGR rate had the greatest effect on the combustion characteristics for both of the tested fuels. According to the HRR results, both fuels produced double peaks of HRR: the first peak reflects the premixed combustion process, and the second peak corresponds to the mixing controlled combustion phase. However, the transition from premixed combustion into mixing controlled combustion became less explicit with an increasing EGR rate for both fuels. In addition, the HRR results indicate that an increasing EGR rate caused a progressive increase in the peak HRR during the premixed burn fraction and shifted the location of the occurrence later toward the expansion stroke. This shift in heat release revealed a delay of the combustion processes due to the prolonged ignition delay. In the present study, the timing difference between the SOI and start of combustion was defined as the ignition delay. The SOI was confirmed from the injector current signal trace and the start of combustion by analyzing the first appearance of positive heat release. As can be seen in Fig. 21, the plot of ignition delay versus EGR rate evidently shows that the increase in the EGR rate caused a progressive increase in the ignition delay for both fuels. In fact, in comparison with the corresponding baseline diesel, the use of the PME fuel resulted in a shorter ignition delay by an average of 0.3° CA. This is credited to the higher cetane number of the PME fuel compared to baseline diesel, thereby resulting in better ignition quality. Subsequently, the effect of prolonging the ignition delay also caused a progressive increase in the peak HRR with a higher EGR rate. As the EGR increases, the in-cylinder fuel air mixture becomes more homogeneous due to a longer ignition delay, which could have allowed a larger fraction of fuel air mixture to burn during the premixed combustion phase. This effect also explained the phenomenon of the premixed combustion process dominating at a high EGR setting. Related to the EGR effects on NO_x emissions, the prolonged ignition delay retards the combustion events toward the expansion stroke, thus promoting more combustible mixtures to burn at a lower temperature. As a result, NO_x formation via the thermal or Zeldovich mechanism can be greatly reduced. Compared to baseline diesel, the peak HRR at the premix combustion stage is found to be consistently lower with PME fuel across all EGR rates. This scenario explained the lower BSNO_x emissions of the PME fuel with respect to baseline diesel at all EGR rates. In addition, the steepness of the HRR curves during the premixed combustion phase decreased with increasing EGR. Again, this gave evidence of reduced reaction rates and decreased BSNO_x with increasing EGR for baseline diesel and the PME fuels.

Fig. 20 In-cylinder pressure, HRR and injector current signal versus crank angle for engine operation with PME (top) and baseline diesel (bottom) at various EGR rates.

Fig. 21 Ignition delay at various EGR rates for engine operation with baseline diesel and PME fuels.

Another important aspect is related to in-cylinder combustion pressure traces. A pronounced change in peak combustion pressure was observed with increasing EGR for both fuels. As shown in Fig. 22, it was found that, in general, regardless of the EGR setting, the PME fuel exhibited a lower peak pressure than baseline diesel owing to the marginal decreases in the HRR during the premixed combustion phase. On average, it is found that PME produced 1.2 bar lower maximum combustion pressure compared to baseline diesel across all EGR rates. The results also indicated that a higher EGR rate tends to lower the maximum cylinder pressure during the expansion stroke. This may be due to the combined effect of greater heat capacity, chemical and thermal effects, with the most significant effect being the dilution effect, which extends the ignition delay duration and thus enhances the in-cylinder charge mixing.⁴⁸ Consequently, the premixed phase of combustion occurred late in the expansion stroke, thereby lowering the peak pressure and reducing BSNO_x.

Fig. 22 Maximum combustion pressure at various EGR rates for engine operation with baseline diesel and PME fuels.

4.3. Strategy for simultaneous BSNO_x–smoke reduction

A vast amount of studies have been carried out to reduce the emissions of pollutants from diesel engines and the attempts are still in progress.^36,49–51 Specifically, the simultaneous reduction of NO_x and smoke is the most challenging aspect in the reduction of diesel emissions. Emission controls at the source level are the most effective techniques in reducing the pollutants of a diesel engine since they are economical when compared to the treatment of exhaust gases.⁵² Injection timing is a key parameter that directly affects the combustion and exhaust emissions. Using EGR for NO_x reduction is another promising approach in diesel engines, but there could be an increase in smoke emissions. Simultaneous reduction of both emission species from the levels of fossil diesel is possible with the use of biodiesels.³² Hence, a comparative analysis of the exhaust BSNO_x and smoke opacity levels between the engine fueled with neat PME biodiesel and that with fossil diesel is discussed in this section to observe the effect of injection timing and EGR variation on BSNO_x and smoke level.

Fig. 23 shows the BSNO_x–smoke plot for PME and baseline diesel with various EGR rates and SOI timings. Generally, the overall trend shows that the BSNO_x emissions for both of the tested fuels reduced with a higher EGR rate and later SOI timing. A substantially lower level of BSNO_x below the EURO V and EURO VI emission standard can be achieved by the late SOI timing of 0° ATDC and at over 45% EGR, respectively. The maximum BSNO_x reduction for diesel operation is achieved with the variation of injection timing and addition of EGR, with a reduction of 72.6% and 97%, respectively. However, both techniques show a penalty on smoke emission as compared to the baseline engine operation without EGR and with an SOI of −11° ATDC. With the PME fuel operation, it is possible to reduce the smoke emission while maintaining a similar reduction in BSNO_x. The results indicate that about 50% and 46% reduction in smoke emission can be attained when the PME biodiesel is coupled with the strategies of late SOI timing and a high EGR level, respectively. Hence, simultaneous BSNO_x and smoke reduction from the levels of fossil diesel is possible with the use of PME biodiesels in parallel with the implementation of late SOI timing or a higher EGR rate in a diesel engine.

Fig. 23 BSNO_x–smoke opacity plot for the tested fuels.

5. Conclusions

In this study, the performance, emission and combustion characteristics of PME have been experimentally investigated in a high-pressure common-rail DI diesel engine. The effect of SOI timing and EGR were investigated at a constant speed of 1500 rpm and a BMEP of 0.4 MPa. The following main conclusions can be drawn from the present study.

1. The physico-chemical properties of the produced PME biodiesel meet the ASTM D6751 standard.

2. The single-cylinder diesel engine had been successfully converted to run with an electronically common-rail fuel injection system using the Arduino microcontroller as an engine ECU. This system was able to offer full control over all injection parameters.

3. Based on the highest BTE and the reasonable NO_x level, the optimum injection timing is found to be at −11° ATDC for both of the baseline diesel and biodiesel operations.

4. The engine-out-responses under higher EGR (>30%) conditions have been investigated in the present study using PME biodiesel and compared with baseline diesel. A substantially lower level of BSNO_x, below the EURO VI emission standard, can be achieved at over 45% EGR.

5. With PME fuel operation, it is possible to reduce the smoke emission while maintaining a similar reduction in BSNO_x. The results indicate that about 46% reduction in smoke emission can be attained when the PME biodiesel is operated under high EGR conditions.

6. Simultaneous BSNO_x and smoke reduction from the levels of fossil diesel is possible with the use of PME biodiesels in parallel with the implementation of late SOI timing or higher EGR rate in diesel engine.

Nomenclature and symbol

AOME	Algal oil methyl ester
ASTM	American society for testing and materials
ATDC	After top dead centre
BMEP	Brake mean effective pressure
BSCO	Brake specific carbon monoxide
BSFC	Brake specific fuel consumption
BSNOx	Brake specific nitrogen oxide
BTE	Brake thermal efficiency
CA	Crank angle
CI	Compression ignition
CO	Carbon monoxide
CO2	Carbon dioxide
DI	Direct injection
ECU	Electronic control unit
EGR	Exhaust gas recirculation
EGT	Exhaust gas temperature
FAME	Fatty acid methyl ester
HC	Hydrocarbon
HRR	Heat release rate
IC	Internal combustion
NOx	Nitrogen oxide
PFI	Port fuel injection
PM	Particulate matter
PME	Palm methyl ester
PWM	Pulse-width-modulation
Rpm	Revolution per minute
SOC	Start of combustion
SOI	Start of injection
TDC	Top dead centre
VOCs	Volatile organic compounds
λ	Relative air–fuel ratio (lambda)

Acknowledgements

The authors would like to acknowledge the Ministry of Higher Education (MOHE) of Malaysia and University of Malaya for financial support through a HIR grant (UM.C/HIR/MOHE/ENG/07) and a Postgraduate Research Grant (PPP) (grant number PG035-2012B).

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