Abstract

The present work emphasizes the effects of injection timing on the characteristics of a 5.2-kW powered four-stroke diesel engine using biogas and its heat loss analysis. The biogas is obtained from food waste consisting of methane (CH4)—88.1% and carbon dioxide (CO2)—11.8% as the composition. The biogas (BG) is selected by mass basis ranging from 20% to 60% with 10% increments and is used to operate the engine by dual-fuel mode. The effect of three injection timings such as 25.5 deg (retarded), 27.5 deg (actual), and 29.5 deg (advanced) before top-dead center (bTDC) under dual-mode operation to enhance the properties of the engine is studied, and the results are compared with diesel mode at actual injection timing. Maximum brake thermal efficiency of 30.1% was observed for BG20 operated at 29.5-deg bTDC injection timing (IT). The dual mode operated at the injection timing of 29.5-deg bTDC showed an increase in cylinder pressure compared to diesel by 11.9% at full load conditions, whereas carbon monoxide emission was lower by 5.2% at 29.5-deg bTDC IT than diesel, and nitrogen oxide emission was lower at 25.5 deg bTDC IT than diesel mode by 45%. Besides, at 75% engine load, the least amount of heat losses was observed for BG50 exhibiting effective conversion of fuel energy into equivalent work higher than that of diesel by 2.2%, respectively.

1 Introduction

The naturally occurring prime energy resources such as fossil fuels, coal, nuclear fuel, and natural gas provide a major contribution to meet the energy demand required for the development of the nation [1]. However, due to their rapid reduction, impact on natural habitats, and harmful effects on the human health, environmental degradation have motivated the researchers and policymakers to focus on the alternative energy resources to replace the depleting fossil fuels and on the same context to meet the energy demand [2]. Biogas (BG) shows a good potential to be one of the renewable energy which is used as prominent energy delivering resource for the prime movers such as diesel, petrol, homogeneous charge compression ignition (HCCI), and dual-fuel engines with necessary adaptations required for power generation and also for other applications [3]. With the lack of oxygen, the higher level organic matter decomposes to produce biogas containing a combination of gases most commonly 50–90% CH4, 22–55% CO2, and remaining traces of other constituents. Compared to other gaseous fuels, biogas is considered one of the prominent alternative fuels for the dual-mode-operated engines due to broader ignition limits, homogeneous mixing nature with other fuels, higher anti-knocking properties, and less emissions, and moreover, it is obtained from abundantly available waste [410]. On the other hand, biogas has its own restrictions for direct utilization in diesel engines because of lower cetane number, lower heating value, as well as higher auto-ignition temperature. Hence, with fewer modifications, biogas can be effectively operated in dual fuel mode in diesel engines [1113]. Dual fuel mode is the best approach to operate any gaseous fuels in diesel engines and provide a better platform for improving engine efficiency by undergoing effective combustion resulting in lower emissions [1416].

However, dual-fuel engine operated with biogas has a few drawbacks such as reduced engine efficiency, higher CO, as well as unburnt hydrocarbons (UBHC) emissions. This is due to the fact that diesel engine operating parameters such as injection timing (IT), injection pressures, and compression ratios (CRs) are optimized based upon the chemical and physical properties of diesel fuel [1720]. As far as dual fuel mode is concerned, the engine is operated without optimizing the operating parameters. Hence, for attaining higher thermal efficiency of dual-fuel engines using biogas, the engine has to be optimized [3,21,2224]. The IT of the diesel engine is a very crucial parameter that influences the engine parameters. As the diesel fuel is injected with varying timings, the condition of the fuel gets varied causing ignition delay to increase. In the advance IT, the temperature and pressure of air at the initial stage will be lower, such that the ignition delay increases. As the injection is retarded, the temperature as well as the pressure of air considerably increases. However, the ignition delay decreases drastically. Therefore, it is understood that IT has a larger impact on the brake thermal efficiency (BTE), brake-specific fuel consumption (BSFC), and nitrogen oxide (NOx) emissions. Hence, there is a necessity to explore the consequence of IT especially during the dual-mode operation using the biogas [25,26].

Numerous researches on the IT effects upon the diesel engine operated with dual fuel mode using various fuels have been reported. Karthic et al. [27] explored the effect of altering the injection pressure and timing in the single-cylinder, constant speed diesel engine using new biodiesel derived from Syzygium cumini oil. The engine was operated for various proportions such as B30, 70, and 100 at injection pressures of 200–260 bar with an increment of 20 and at IT of 21–25 deg before top-dead center (bTDC) with 2-deg increment. The best performance was attained at advanced IT 25-deg bTDC and at 240 bar injection pressure such that BTE increased by 17.8% and 16.6% for diesel and B30, respectively. The influence of port fuel injection (PFI) timing on direct injection (DI) engine with six cylinders under dual fuel mode condition with biogas was analyzed by Mohand et al. [28] using GT-power simulation. Major differences were observed for injecting outside of the intake window such that after injecting outside the intake stoke resulted in awaiting of the gas at the intake port. By observing both the simulation and test results, the authors concluded that cylinder 1 showed lean in nature whereas cylinder 6 showed enrich behavior which is due to the improved combustion of methane beneath richer conditions. Kannan et al. [29] explored the potential of IT and pressure in a diesel engine using diestrol consisting of 30% waste cooking oil of palm oil methyl ester, 60% diesel, and 10% ethanol. At the 25.5-deg bTDC IT and 240 bar injection pressure, 31.3% of higher BTE was achieved such that lower CO, SO (smoke opacity), and CO2 were observed for diestrol. Compared to diesel, diestrol showed 4.3% lower NOx emission but alternatively, UBHC levels were increased. Kumar et al. [25] demonstrated the behavior of diesel engine fuelled with 20% tamarind seed biodiesel when it is subjected to advanced (27-deg bTDC), actual (23-deg bTDC), and retarded (19-deg bTDC) ITs. The authors found that the maximum BTE was enhanced by 3.1% with the reduction in CO by 17.3%, NOx by 31.3%, and SO by 8.1% for full load at retarded IT compared to the actual IT of the engine.

1.1 Objectives of the Research Work.

The preliminary aim of the current investigation is to effectively utilize the biogas obtained from food waste as a renewable fuel to operate the diesel engine under dual fuel mode. Further, to examine the engine characteristics under the influence of three different ITs namely Advanced (29.5-deg bTDC), Actual (27.5-deg bTDC), and Retarded (25.5-deg bTDC) using different proportions of raw biogas ranging from BG20 (i.e., BG20—20% of biogas by mass basis) to BG60 with 10% increment operated and to compare the outcomes with fossil fuel at actual IT. In addition to this, the various forms of heat losses occurring during the engine operation under diesel and dual-fuel mode at actual IT are analyzed. Moreover, the study provides a path for the researchers to design the engine meant to operate with gaseous fuels that are very vulnerable to changes based upon varying the ITs. Besides, no major work has been reported in providing the IT effects and heat loss analysis combination for higher biogas proportions.

1.2 Materials Used.

For the present investigation, the diesel engine is driven with the biogas acquired from the biogas plant which is in charge of the city corporation and the plant is located 14.1 km away from the National Institute of Technology Karnataka (NITK) campus. The corporation collects the food waste produced by different hotels located inside the town and generates the biogas from the anaerobic digestion plant. The plant is a two-tonne capacity that generates 140–180 m3 of biogas, and as a byproduct, 120–150 kg of compost is created daily. The composition of the biogas was measured using gas chromatography [15], and the properties of different fuels obtained are given in Table 1. The raw biogas is used in the present investigation, and for comparison purposes, the property of biogas is shown in the table and is obtained from the literature.

Table 1

Properties of different fuels [10]

PropertyDieselBiogasRaw biogas
Composition (%)C12H26CH4—63.74%
CO2—36.26%
CH4—88.10%
CO2—11.89%a
Lower heating value (MJ/kg)42.06b26.4228.05#
Density (kg/m3)8401.260.75
Auto-ignition temperature (°C)275644
PropertyDieselBiogasRaw biogas
Composition (%)C12H26CH4—63.74%
CO2—36.26%
CH4—88.10%
CO2—11.89%a
Lower heating value (MJ/kg)42.06b26.4228.05#
Density (kg/m3)8401.260.75
Auto-ignition temperature (°C)275644
a

Tested values.

b

Calculated values.

2 Experimental Setup

The investigation is executed on a computerized diesel engine test rig which comprises four-stroke, single-cylinder, naturally aspirated, water-cooled and direct injection engine located in the Internal combustion (IC) engine research laboratory, inside the institute. Using an eddy current dynamometer, different loads can be applied to the engine by means of controlling the voltage with the current supply simultaneously. In addition to this, the engine is controlled and operated by providing the inputs through the control panel. Using the IC engine soft version 9.0, various parameters such as air flow, fuel consumption, mass flowrate of fuel, brake power (BP), indicated mean effective pressure, mechanical efficiency, fuel line pressure, torque, mean gas temperature, and net heat release rate are acquired, and the data are saved in the computer. Complete technical details of the modified diesel engine and instruments used for the investigation purpose are specified in Table 2.

Table 2

Technical specifications of the engine test rig and instruments used

Engine categoryFour-stroke, single-cylinder, constant speed, water-cooled diesel engine
ModelKirloskar TV1
Brake power5.2 kW @ 1500 rpm
Bore × Stroke87.5 × 110 mm
Compression ratio17.5:1
Length of the connecting rod234 mm
Type of the dynamometerEddy current
Propeller shaftWith universal joints
Load measurementStrain gauge load cell
Piezo powering unitCuadra, AX-409
Fuel and air measurementDifferential pressure transmitter
Crank angle sensorResolution 1 deg @ 5500 rpm with TDC pulse
Data acquisition devicePCI-1050
Dynamometer loading unitApex, Model: AX-155. Type constant speed, Supply 230 V AC
Engine categoryFour-stroke, single-cylinder, constant speed, water-cooled diesel engine
ModelKirloskar TV1
Brake power5.2 kW @ 1500 rpm
Bore × Stroke87.5 × 110 mm
Compression ratio17.5:1
Length of the connecting rod234 mm
Type of the dynamometerEddy current
Propeller shaftWith universal joints
Load measurementStrain gauge load cell
Piezo powering unitCuadra, AX-409
Fuel and air measurementDifferential pressure transmitter
Crank angle sensorResolution 1 deg @ 5500 rpm with TDC pulse
Data acquisition devicePCI-1050
Dynamometer loading unitApex, Model: AX-155. Type constant speed, Supply 230 V AC

A pressure transducer (PCB Piezotronics, Model: HSM111A22) is attached to the cylinder head for measuring the maximum peak cylinder pressure, and Chromel_Alumel made K type thermocouples are utilized to determine the temperatures across several positions of the engine test rig. Signals in analog form coming from the engine are translated into the digital signal by data acquisition system (DAQ) PCI-1050, and all the data are stored in the computer. By means of an exhaust gas analyzer (EGA), the emissions from the engine including O2, CO, NOx, and HC are determined. Table 3 provides the in-detail description of the gas analyzer used in the current study.

Table 3

Details of the exhaust gas analyzer (AVL)

Measuring factorsRange of the measurementResolution
Carbon dioxide (CO2)0–20% vol0.1% vol
Hydrocarbons (HC)0–20,000 ppm10 ppm
Carbon monoxide (CO)0–10% vol0.01% vol
Nitrogen oxide (NOx)0–5000 ppm1 ppm
Lambda (λ)0–9.9990.001
Oxygen (O2)0–22% vol0.01% vol
Measuring factorsRange of the measurementResolution
Carbon dioxide (CO2)0–20% vol0.1% vol
Hydrocarbons (HC)0–20,000 ppm10 ppm
Carbon monoxide (CO)0–10% vol0.01% vol
Nitrogen oxide (NOx)0–5000 ppm1 ppm
Lambda (λ)0–9.9990.001
Oxygen (O2)0–22% vol0.01% vol

A cylinder containing compressed biogas is fixed to the engine such that a check valve is provided to avoid the backpressure during the induction of gas. The pressure of the biogas is regulated by a pressure regulator, and the amount of biogas to be inducted is controlled by flowmeter [16]. A schematic block diagram of the experimental setup is represented in Fig. 1.

Fig. 1
Schematic block diagram of the engine test rig
Fig. 1
Schematic block diagram of the engine test rig
Close modal

2.1 Heat Loss Analysis Calculations.

Using a calorimeter, the heat loss occurring in the premises of the engine after the combustion using diesel as well as biogas proportion is analyzed in this work by determining the temperature variations at different positions across the engine setup using six different K-type thermocouples. Rotameter is utilized to determine the flowrate of water entering and leaving the engine such that water from the engine jacket is used as a working medium inside the calorimeter. Using the steady-state flow energy expression from the thermodynamics for a control volume system heat loss distribution [30] is calculated as mentioned below:

Heat equivalent of work (%):
QEW=BPmf×CV
(1)
where QEW is the heat equivalent of work, BP is the brake power in kW, mf is the mass flowrate of the fuel in kg/h, and CV is the calorific value in MJ/kg.
Heat by jacket cooling water (%):
QJCW=mw×Cpw×(T2T1)mf×CV
(2)
QJCW is the heat loss by jacket cooling water, mw is the mass flowrate of water in kg/h, Cpw is the specific heat of water in kJ/kg-K, T1 is the temperature of the water inlet in K, and T2 is the temperature of the water outlet in K.
Heat by exhaust gas (%):
Qg=(ma+mf)×Cpg×(T5Tatm)mf×CV
(3)
Qg is the heat loss by exhaust gas and ma is the mass flow rate of intake air in kg/h; Cpg is the specific heat of exhaust gas in kJ/kg-K, T5 is the exhaust gas to calorimeter inlet temperature in K, and Tatm ambient temperature in K.
Unaccounted heat (%):
UAH=100(QEW+QJCW+Qg)
(4)
UAH is the unaccounted heat loss.

2.2 Methodology.

In this present work, the investigation executed by delivering the compressed biogas at the inlet manifold of the engine with different fractions of biogas ranging from BG20 to BG60 with a 10% increment is performed through dual mode (DM). By Eq. (5), the biogas flowrate to be inducted into the engine is determined as given below [16]:
mf=BSEC×BPCV
(5)
where BP is the brake power in kW, BSEC is the brake-specific energy consumption in MJ/kW-h, and CV is the calorific value of the fuels in MJ/kg.

The biogas fractions for every load will be changing and are obtained by determining the BSEC of the diesel fuel considering one particular load using Eq. (5). After finding the BSEC value of diesel, the amount of biogas fractions to be inducted is obtained in the next step. For example, the mass flowrate of BG20 is obtained by finding out the product of 20% biogas and BSEC of diesel fuel for the specified load. As the corresponding BP and CV are already known, by substituting the determined values, the mass flowrate of BG20 is obtained for the respective load. Similarly, the mass flowrate for different biogas fractions is determined by using the same Eq. (5) for every load by the same procedure. Induction of BG20 at the intake manifold is achieved by releasing the gas gradually by the mass flowmeter; the required operating pressure of the biogas is achieved via governing the pressure regulator. As the mass flowrate of biogas is increased, the engine speed will increase; as a result, the engine speed is reduced and operated at a constant speed of 1500 rpm by regulating the flowrate of the diesel fuel through the speed governor, and hence, the consumption of diesel is reduced. Inducted biogas mixes with the intake air inside the air filter homogeneously and flows into the combustion chamber; thereby, additional power is created after the combustion takes place. In a similar manner, the experiment is performed for the remaining gases at all the loads.

The engine is operated for three different ITs namely Advanced (29.5-deg bTDC), Actual (27.5-deg bTDC), and Retarded (25.5-deg bTDC), respectively. The required IT is achieved by changing a small part called the shim which is located beneath the fuel pump such that each shim is of 2 mm thickness and 2-deg CA. The actual IT of the diesel engine is 27.5-deg bTDC and consists of two shims; if 1 shim is removed, then the IT will be advanced to 29.5-deg bTDC, and if all the three shims are included, then the IT will be retarded to 25.5-deg bTDC. The experiment is conducted for all three ITs for the BG20–BG60 proportion, which are inducted at the intake manifold of the engine operated for 25–100% load with 25% increment at a constant speed of 1500 rpm. The different loading is achieved from 0–100% load by varying the voltage and current knob simultaneously until the engine speed arrives at a constant speed of 1500 rpm. However, in the case of biogas fractions, the zero load is omitted and is directly started from 25% load until the final load. The test is conducted first for diesel only at the actual timing (27.5-deg bTDC), and then, the process is switched over to the dual mode by inducting the biogas which is operated for the three ITs. Thereafter, the results of the engine performance using biogas for three ITs are compared with diesel at actual IT. Figure 2 depicts the experimental arrangements made; Fig. 2(a) illustrates the modification of the diesel engine and the method of inducing the biogas into engine through dual fuel mode, whereas Fig. 2(b) represents the control panel connected to the engine and the matrix of the experimental work is clearly mentioned in Table 4.

Fig. 2
Experimental setup: (a) modified dual diesel engine and (b) control panel connected to the engine
Fig. 2
Experimental setup: (a) modified dual diesel engine and (b) control panel connected to the engine
Close modal
Table 4

Matrix of the experiment

MethodFuel usedIT (bTDC)RatiosLoadSpeed (rpm)
Diesel modeOnly diesel27.5 deg100%0%, 25%, 50%, 75%, 100%1500
Dual modeRaw biogas-diesel25.5 deg (Retarded)
27.5 deg (Actual)
29.5 deg (Advanced)
20% to 60% (10% increment)25%, 50%, 75%, 100%
MethodFuel usedIT (bTDC)RatiosLoadSpeed (rpm)
Diesel modeOnly diesel27.5 deg100%0%, 25%, 50%, 75%, 100%1500
Dual modeRaw biogas-diesel25.5 deg (Retarded)
27.5 deg (Actual)
29.5 deg (Advanced)
20% to 60% (10% increment)25%, 50%, 75%, 100%

2.3 Uncertainty Analysis.

Uncertainty provides the complete details of errors associated with the experimental results or instruments by means of the partial differential method. Uncertainty is caused because the environment, selection of instrument, calibration, observation, human errors, and experimental procedure [31]. In this work, the uncertainty is analyzed for the dependent variables such as BP, BSEC, and BTE and is determined by a partial differential method with uncertainty associated for every instrument used as shown in Table 5. By calculating the standard deviation, mean, and error, the uncertainties for independent parameter considered for an overall experiment is obtained.

Table 5

List of the uncertainties of different parameters

MeasurementUncertainty (%)
BP0.55
BTE0.94
BSEC0.85
Load0.92
Speed0.61
CO0.41
UBHC0.74
NOx0.65
Sootness0.87
Cylinder pressure0.74
MeasurementUncertainty (%)
BP0.55
BTE0.94
BSEC0.85
Load0.92
Speed0.61
CO0.41
UBHC0.74
NOx0.65
Sootness0.87
Cylinder pressure0.74
The total uncertainty determined is as follows:
=(BP)2+(BTE)2+(BSEC)2+(Load)2+(speed)2+(CO)2+(UBHC)2+(NOx)2+(SO)2+(CP)2=(0.55)2+(0.94)2+(0.85)2+(0.92)2+(0.61)2+(0.41)2+(0.74)2+(0.65)2+(0.87)2+(0.74)2=±2.37%

3 Results and Discussions

3.1 Brake Thermal Efficiency.

The BTE of the engine using biogas proportion ranging from BG20 to BG60 operated at the three ITs 29.5-deg, 27.5-deg, and 25.5-deg bTDC, along with the diesel at 27.5-deg bTDC operated for all load is shown in Figs. 3(a)3(c). The engine operated under dual mode at all the loads using different proportions of biogas showed lower BTE for all the three ITs. The performance of the engine is evaluated and equated with diesel at 27.5-deg bTDC IT. The diesel indicated higher BTE compared to dual mode for the respective load. As far as the different IT is concerned, the maximum BTE was observed at advanced (29.5-deg bTDC) IT whereas retarded (25.5 deg bTDC) IT showed lower BTE. This decline in BTE is because of several factors such as the lower calorific value of biogas, lower flame propagation, and induction of a higher rate of biogas–air mixture causing negative compression work which leads to higher consumption of the biogas–air fuel mixture [3133].

Fig. 3
(a)–(c) Influence of injection timing on BTE for dual mode
Fig. 3
(a)–(c) Influence of injection timing on BTE for dual mode
Close modal
Brake thermal efficiency of the engine operated under the dual mode is determined by the equation:
[ηth]DM=BP(mD*×LHVD*)+(mBG*×LHVBG*)×100
(6)
where [ηth]DM is the brake thermal efficiency (%) of dual fuel mode, BP is the brake power (kW), mD* is the mass flowrate of diesel (kg/h), mBG* is the mass flowrate of biogas, LHVD* is the lower heating value of diesel and LHVBG* is the lower heating value of biogas (MJ/kg) [16]. BG20 operated at advanced IT showed maximum BTE of 30.1% at full load, and BG20 showed a lower BTE of 4.9% at full load compared to diesel, whereas BG60 exhibited a lower BTE of 25.9% and is 28.4% lower than diesel at full load for IT of 29.5-deg bTDC. The trend of reduction in BTE is instigated due to the induction of biogas at the intake manifold. As a result, intake of air decreases causing a reduction in volumetric efficiency. As observed from Figs. 3(a)3(c), advanced IT shows maximum BTE compared to actual and retarded IT at all the load for every biogas proportion ranging from BG20 to BG60. This is because at the advanced timing additional period is accessible for homogeneous mixing of the air-fuel as compared to actual and retarded timings and also due to the combination of higher vaporization of the diesel fuel as well as efficient combustion of biogas–air mixture generating additional power [11,31].

3.2 Cylinder Pressure (P-θ).

The variations of peak cylinder pressure for 100 cycles of both diesel as well as dual fuel operated at full load under the influence of three ITs are shown in Figs. 4(a)4(c). Generally, it is observed that as the biogas proportion increases the cylinder pressure increases more than the diesel. BG60 shows higher cylinder pressure of 60.4 bar, operated at 27.5 deg bTDC, whereas the cylinder pressure of BG20, BG30, BG40, and BG50 is 57.9 bar, 58.7 bar, 59.5 bar, and 60 bar correspondingly. This is because of the dilution of oxygen concentration as a result of biogas induction at the intake manifold such that volumetric efficiency gradually decreases. This development at the premixed combustion phase causes ID to get elongated creating peak cylinder pressure to increase at a higher level [15,32].

Fig. 4
(a)–(c) Influence of injection timings on peak cylinder pressure for dual mode
Fig. 4
(a)–(c) Influence of injection timings on peak cylinder pressure for dual mode
Close modal

Advanced IT (29.5-deg bTDC) showed maximum cylinder pressure for all the biogas proportion compared to actual and retarded IT operated at full load. This is because, as compared to actual and retarded IT, enhanced homogeneous mixing of air–biogas was achieved within the cylinder and due to the development of fuel-rich mixture within the cylinder. As a result, at the premixed combustion phase, quick burning of air-biogas takes place causing combustion to appear at the earliest such that a large quantity of dual fuel gets burned before the piston reaches the top dead center (TDC) causing the cylinder pressure to reach maximum and appears closer toward TDC [31,34]. BG60 showed a maximum cylinder pressure of 64.1 bar, whereas the cylinder pressure for BG20 is 60.9 bar at advanced IT. Compared to diesel, BG60 showed an increase in cylinder pressure by 11.9% at 29.5-deg bTDC. On the other hand, cylinder pressure for BG20, BG30, BG40, and BG50 increased by 7.3%, 7.9%, 9.2%, and 11% compared to diesel, respectively.

3.3 Net Heat Release Rate.

Net heat release rate (NHRR) for the diesel as well as dual mode is described in Figs. 5(a)5(c) for the three ITs operated at full load. Basically, the concept of NHRR is a difficult process, as it includes the combustion of binary fuels having different properties. NHRR essentially depends on the quality of the inducted air–biogas mixture along with the diesel and also on parameters such as energy density and mass flowrate of both the diesel and biogas fuel which depict the intensity of heat release rate. Under dual mode, HRR is formed due to the three levels of the combustion process namely combustion of diesel fuel, combustion of biogas, and the pre-ignition at the consequent flame propagation [31,33]. As observed from Figs. 5(a)5(c) the NHRR is higher for dual mode compared to diesel, such that the rate of heat release occurring at the premixed combustion stage is influenced by the rate of combustion at the starting stage, formation of the mixture, and ignition delay [16,35]. Moreover, due to the induction of biogas, there is a delay in the appearance of NHRR, caused because of CO2 presence and its higher specific heat.

Fig. 5
(a)–(c) Influence of injection timing on Net heat release rate for dual mode
Fig. 5
(a)–(c) Influence of injection timing on Net heat release rate for dual mode
Close modal
For dual mode, as the IT is advanced, the NHRR appears at the early stage which provides more platform for full spray and quantity of the fuel mixture. Because of more time available, the combustion initiates at the earliest burning more amount of fuel and energy is released at a higher rate [14,31]. Based on the correlations related to the first law of thermodynamics, the NHRR for dual fuel mode at every crank angle is determined from the following expression:
Q=W+dudt
(7)
where Q′ is the net heat generated (J) and heat transfer taking place at the cylinder wall and W is the rate of work done (J) indicating the displacement done by the system. Equation (7) is further modified as per the ideal gas condition for the unit mass system as shown below:
Q=[CvR+1]×P×dvdt+CvR×V×dPdt
(8)
Equation (8) is further modified in terms of ideal gas condition, and time (t) is replaced with crank angle (θ) and CP/CV the final expression is as follows:
Q=γ*γ*1×P×[dvdθ]+1γ*1×V×[dpdθ]
(9)
where γ* is the specific heat ratio, P′ is the instant cylinder pressure expressed in N/m2, and V′ is the cylinder volume in m3 and is determined by the values of the crank angle and geometry of the engine. Eventually, the NHRR under dual-mode operation completely depends on the fuel quality of the biogas, air, and diesel combination, the mass flowrate as well as the LHV of the fuels and NHRR is the net heat release rate measured in J/°CA [31,34]. Maximum NHRR is observed for BG60 with 33.5 J/°CA and shows an increment in NHRR compared to diesel by 29.1% at advanced IT operated under full load, whereas other biogas proportions BG20, BG30, BG40, and BG50 showed NHRR of 27.3 J/°CA, 28 J/°CA, 29 J/°CA, and 30.9 J/°CA, respectively.

3.4 Carbon Monoxide.

Variations in CO for diesel as well as dual mode operated for the defined ITs and load are represented in Figs. 6(a)6(c). Among the diesel and dual fuels, CO emission was found to be higher for dual fuels due to the lower flame propagation, partial combustion, lower intake temperature of the air–gas mixture, higher equivalence ratio, and most important reason is the appropriate time available for complete combustion to take place. On the other hand, the existence of 11.8% CO2 in biogas causes incomplete combustion by diluting the air–biogas mixture [34,36]. However, the flame generated from the combustion of diesel fuel is generally retained in the flame zone until the minimum value of self-ignition is attained by the biogas–air mixture, thus promoting incomplete combustion. In the case of IT's, the dual mode showed a decrement in CO emissions as the IT is advanced toward 29.5-deg bTDC, which is due to the maximum temperature attained inside the cylinder and improvement in the oxidation reaction taking place between the gas molecules [35,37].

Fig. 6
(a)–c) Influence of injection timing on CO emissions for dual mode
Fig. 6
(a)–c) Influence of injection timing on CO emissions for dual mode
Close modal

The retarded IT showed higher CO emissions as it underwent a poor combustion process due to a lack of oxygen and less time for complete combustion to take place. Among the three ITs, advanced IT showed lower CO emission as compared to actual and retarded IT for dual-mode operation [17,31]. As observed from the figure, BG20 showed lower CO emissions, whereas BG60 showed higher CO emissions for three ITs. BG60 showed CO emissions higher by 20% and 11.7% at 3/4th and full load compared to diesel mode, whereas BG20 emitted lower CO emissions by 2.5% and 5.2% at 3/4th and full load compared to diesel, respectively.

3.5 Nitrogen Oxide.

The influence of IT's on NOx emissions emitted from the diesel engine operated in diesel and dual mode are shown in Figs. 7(a)7(c). In general, the NOx formation is analyzed from zeldovich mechanism which indicates factors such as high temperature of the combustion chamber, presence of adequate air concentration, retention time, and nitrogen molecules which are responsible for NOx formation [15,33].

Fig. 7
(a)–(c) Influence of injection timing on NOx emissions for dual mode
Fig. 7
(a)–(c) Influence of injection timing on NOx emissions for dual mode
Close modal

Eventually, in the dual-mode operation, substantial decrease in NOx emission compared to diesel mode is witnessed because, as the biogas is inducted at the intake manifold, the volumetric efficiency decreases as a result of lack of oxygen concentration. However, due to the higher equivalence ratio, the NOx emission increased as the load increased for both diesel and dual mode. The other reasons for the NOx emission to decrease are because of CO2 presence in the biogas, which acts as a diluent and due to its higher molar specific heat; thereby, the maximum in-cylinder temperature decreases [35]. In the case of the three ITs, retarded IT of 25.5-deg bTDC showed less NOx emission with respect to diesel mode, whereas an increase in pressure and temperature is observed in the combustion chamber operated at advanced IT for different proportions of biogas; this development is because of early induction of the diesel, such that additional interval is available for homogeneous mixing with biogas–air mixture and complete combustion to take place [31,33]. The retarded IT of 25.5-deg bTDC with BG60 emitted lower NOx emissions than diesel at actual ITs by 52% and 45% at 75% and 100% load, respectively, whereas BG20 exhibited higher emissions among the other biogas proportions at 25.5-deg bTDC but showed lower NOx emissions compared to diesel by 23.4% at full load. Even though advanced IT showed higher emissions than other ITs, less NOx emission than diesel at actual timing was observed by 39.8% and 27.8% at three-fourth and full load, respectively.

3.6 Smoke Opacity.

Variations of smoke opacity or sootness for diesel and dual-fuel modes for the three ITs operated at various loads are depicted in Figs. 8(a)8(c). The engine operated using biogas shows discrete advantage over diesel mode with the lower emission of sootness. Conversely, sootness directly depends on the fuel-rich region occurring at the diffusion stage due to the incomplete combustion and on the fuel consumption [23]. As it is well known, diesel consists of higher aromatic compounds which are responsible for higher SO emissions. These kinds of compounds are absent in the biogas; instead, it contains higher methane composition and has less affinity toward soot formation [35]. Figures 8(a)8(c) specifies that as the load is increased the SO emission increases in the case of dual fuel mode but less than diesel. The reason behind this development is due to the increase in ID, such that the higher oxidation of soot particles formed is suppressed by the CO2 existence in the biogas, and as a result, the flame front temperature decreases. Diesel mode operated at 27.5-deg bTDC showed higher soot emissions at all load and other ITs compared to dual-mode operation.

Fig. 8
(a)–(c) Influence of injection timing on sootness for dual mode
Fig. 8
(a)–(c) Influence of injection timing on sootness for dual mode
Close modal

Moreover, lower soot emissions were observed for the IT 29.5-deg bTDC compared to diesel and other IT's which is caused because of superior combustion rate, additional time available for oxidation of soot particles inside the cylinder [31,36]. However, the accumulation of additional fuel within the cylinder facilitates proper mixing of air–biogas mixture, and thereby, the reaction rate increases to enrich the oxidation process leading to the formation of less soot emission [24]. BG60 showed less SO emission compared to diesel by 45.6% and 41.3%, respectively, at three-fourth and full load with 29.5-deg bTDC IT. On the other hand, BG20 showed higher SO emissions among other biogas proportions but less than diesel by 20.9% and 18.6%, respectively, at three-fourth and full load.

3.7 Exhaust Gas Temperature.

The exhaust gas temperature (EGT) variations for diesel and different biogas proportions operated at actual IT as well as for varying loads are depicted in Fig. 9. Maximum EGT is observed for diesel fuel indicating the requirement of more heat as the diesel fuel is denser to initiate the combustion process [31]. Moreover, the heavier droplets remain unburnt during the combustion and instead start burning during the after-burning stage. However, a typical development is observed in the case of biogas proportion which depicts lower EGT as compared to diesel fuel caused because of dilution creation in the diesel–biogas mixture by the presence of CO2 in biogas. Conversely, the reason for the drop in EGT is because lower flame propagation of biogas promotes incomplete combustion [35]. Also, it is well known that biogas possesses higher auto-ignition temperature, as a result, more heat energy is absorbed causing the temperature of local flame to decrease during the combustion process such that EGT decreases drastically as the biogas proportion increases [34]. As observed from the graph, every biogas proportion depicted lower EGT compared to diesel for all the operating loads such that BG20, BG30, BG40, BG50, and BG60 exhibited lower EGT than diesel by 3.2%, 6.4%, 10.3%, 14.2%, and 17.4% at 3/4th load and 3.2%, 6.4%, 8.5%, 11.4%, and 14.4%, respectively, for full load at actual IT.

Fig. 9
Exhaust gas temperature variations for different fuels at the standard injection timing
Fig. 9
Exhaust gas temperature variations for different fuels at the standard injection timing
Close modal

3.8 Heat Loss Analysis for Diesel and Dual Mode.

It is evident from the various research surveys that a major portion of the energy liberated from the fuel is utilized in the frictional work especially during a lower load operation. Hence, in this work, heat loss analysis is performed for diesel, BG20 to BG60 at actual IT and higher loads (75% and 100% load). The analysis of the heat equivalent of work, heat by jacket cooling water, heat by exhaust gas, and unaccounted heat losses are determined from Eqs. (1)(5), and the heat losses for every fuel for 75% and full load are illustrated in Table 6. As the biogas flowrate is increased from BG20 to BG60, the unaccounted heat loss is observed to be lower compared to diesel for both 75% and full load, which is due to the lower EGT under dual-mode operation and due to the presence of CO2 [31,35]. The heat loss distribution under uncounted heat for BG20 is found to be 5.4% and 3.1% lower than diesel at 75% and full load. The same reason is applicable to jacket cooling and exhaust gas losses for biogas proportion indicating the major portion of the energy from the biogas combustion is effectively converted into brake power. As explained in the trend of exhaust gas temperature major energy liberated during the combustion is consumed to ignite the biogas as it possesses higher auto-ignition temperature causing the lesser distribution of heat across the engine premises [37]. The heat loss by jacket cooling water for BG60 is higher than diesel by 24.3% and 18.1% at 75% and full load, whereas the heat loss by exhaust gas for BG50 is lower than diesel by 1.7% and 16.7% at 75% and 100% load.

Table 6

Heat loss distribution for diesel mode and dual mode at 75% and 100% load

FuelsHeat equivalent of work (%)Heat by jacket cooling water (%)Heat by exhaust (%)Unaccounted heat (%)
75% L100% L75% L100% L75% L100% L75% L100% L
Diesel27.230.219.920.323.123.429.826.1
BG202626.623.623.522.224.628.225.3
BG3026.427.222.224.624.323.127.125.1
BG4026.921.221.223.224.422.627.526.7
BG5027.828.523.925.322.719.525.626.7
BG6025.827.226.324.824.521.723.426.3
FuelsHeat equivalent of work (%)Heat by jacket cooling water (%)Heat by exhaust (%)Unaccounted heat (%)
75% L100% L75% L100% L75% L100% L75% L100% L
Diesel27.230.219.920.323.123.429.826.1
BG202626.623.623.522.224.628.225.3
BG3026.427.222.224.624.323.127.125.1
BG4026.921.221.223.224.422.627.526.7
BG5027.828.523.925.322.719.525.626.7
BG6025.827.226.324.824.521.723.426.3

Eventually, the heat equivalent of work is higher for diesel at both the loads as the diesel possesses a higher calorific value but conversely the trend is different for dual mode with lower heat equivalence of work because of the LHV. Also, the CO2 presence promotes lower flame propagation indicating less efficiency to convert the energy liberated from the combustion of fuel to equivalent work [17]. The maximum equivalence of the work attained for the diesel at 75% and full load is 27.2% and 30.2%, whereas all the biogas proportion exhibited minimum equivalence of the work compared to diesel for both the loads. For the dual mode, the maximum equivalence of the work is observed to be for BG50 with 27.8% and 28.5% at 75% and full load, respectively. Conversely, BG50 showed effective conversion of work higher than diesel by 2.2% at 75% load and lower by 5.6% at full load, alternatively BG40 showed lower equivalent work than diesel by 1.1% and 8.9% operated for 75% and full load, respectively. From the analysis of heat losses depicted in Table 6 for every test fuels, the effective conversion of fuel energy into equivalent work with minimum heat losses is achieved for BG50 at 75% load showing highest conversion of fuel energy into effective work compared to other biogas proportions.

4 Conclusions

Based on the experimental outcomes of the present investigation using biogas operated in dual mode diesel engine with different ITs, the following conclusions are formulated:

  • Maximum BTE was observed for BG20 with the IT 29.5-deg bTDC by 7.3% higher at full load compared to other ITs but lower efficiency was achieved compared to diesel by 4.9% at IT 27.5-deg bTDC for the same operating load.

  • At 29.5-deg bTDC IT, BG60 showed a maximum cylinder pressure of 64.15 bar and increased cylinder pressure compared to diesel by 11.9% at full load.

  • Highest NHRR is observed for dual mode with 33.5 J/°CA and shows increment in NHRR compared to diesel by 29.1% operated at 29.5-deg bTDC IT.

  • Higher biogas proportion emitted lower NOx emissions at 25.5 deg bTDC IT than diesel mode by 45% at full load. The smoke opacity was observed to be lower for dual mode compared to diesel by 41.3% at full load with 29.5 deg bTDC IT.

  • The unaccounted heat loss decreases for BG20 which is found to be 5.4% and 3.1% lower than diesel at 75% and 100% load, while BG40 showed lower unaccounted heat loss than diesel by 7.7% at 75% load and higher by 2.3% at full load for 27.5-deg bTDC IT.

  • Under dual mode, the maximum equivalence of the work is attained for BG50 with 27.8% and 28.5% at 75% and full load by exhibiting effective conversion of work higher than diesel by 2.2% at 75% load and lower by 5.6% at full load; alternatively, BG40 showed lower equivalent work than diesel by 1.1% and 8.9% operated for 75% and full load for 27.5-deg bTDC IT, respectively.

Conflict of Interest

There are no conflicts of interest.

Data Availability Statement

The authors attest that all data for this study are included in the paper.

Nomenclature

Pmax =

maximum cylinder pressure

mD* =

mass flowrate of diesel

mBG* =

mass flowrate of biogas

rpm =

revolutions per minute

CRDi =

common rail direct injection

Pθ =

cylinder pressure v/s crank angle

SM =

smoke meter

References

1.
Ayhan
,
V.
,
Çangal
,
Ç
,
Cesur
,
İ
,
Çoban
,
A.
,
Ergen
,
G.
,
Çay
,
Y.
,
Kolip
,
A.
, and
Özsert
,
İ
,
2020
, “
Opti of the Factors Affecting Performance and Emissions in a Diesel Engine Using Biodiesel and EGR With Taguchi Method
,”
Fuel
,
261
, p.
116371
.
2.
Sadiq
,
Y. R.
, and
Iyer
,
R. C.
,
2020
, “
Experimental Investigations on the Influence of Compression Ratio and Piston Crown Geometry on the Performance of Biogas Fuelled Small Spark Ignition Engine
,”
Renew. Energy
,
146
, pp.
997
1009
.
3.
Berlini
,
R.
,
Molina
,
R.
,
Hernández
,
J. J.
,
César
,
A.
,
Malaquias
,
T.
,
Coronado
,
C. J. R.
,
José
,
F.
, and
Pujatti
,
P.
,
2020
, “
Experimental Investigation on the Potential of Biogas/Ethanol Dual-Fuel Spark-Ignition Engine for Power Generation: Combustion, Performance and Pollutant Emission Analysis
,”
Appl. Energy
,
261
, p.
114438
.
4.
Akkouche
,
N.
,
Loubar
,
K.
,
Nepveu
,
F.
,
El
,
M.
, and
Kadi
,
A.
,
2020
, “
Micro-combined Heat and Power Using Dual Fuel Engine and Biogas From Discontinuous Anaerobic Digestion
,”
Energy Convers. Manag.
,
205
, p.
112407
.
5.
Mangesh
,
V. L.
,
Padmanabhan
,
S.
,
Tamizhdurai
,
P.
, and
Ramesh
,
A.
,
2020
, “
Experimental Investigation to Identify the Type of Waste Plastic Pyrolysis Oil Suitable for Conversion to Diesel Engine Fuel
,”
J. Clean. Prod.
,
246
, p.
119066
.
6.
Uusitalo
,
A.
,
Uusitalo
,
V.
,
Grönman
,
A.
,
Luoranen
,
M.
, and
Jaatinen-Värri
,
A.
,
2016
, “
Greenhouse Gas Reduction Potential by Producing Electricity From Biogas Engine Waste Heat Using Organic Rankine Cycle
,”
J. Clean. Prod.
,
127
, pp.
399
405
.
7.
Kozina
,
A.
,
Radica
,
G.
, and
Nižetić
,
S.
,
2020
, “
Analysis of Methods Towards Reduction of Harmful Pollutants From Diesel Engines
,”
J. Clean. Prod.
,
262
, p.
121105
.
8.
Makareviciene
,
V.
,
Sendzikiene
,
E.
,
Pukalskas
,
S.
,
Rimkus
,
A.
, and
Vegneris
,
R.
,
2013
, “
Performance and Emission Characteristics of Biogas Used in Diesel Engine Operation
,”
Energy Convers. Manag.
,
75
, pp.
224
233
.
9.
Muralidharan
,
K.
,
Vasudevan
,
D.
, and
Sheeba
,
K. N.
,
2011
, “
Performance, Emission and Combustion Characteristics of Biodiesel Fuelled Variable Compression Ratio Engine
,”
Energy
,
36
(
8
), pp.
5385
5393
.
10.
Chandra
,
R.
,
2011
, “
Performance Evaluation of a Constant Speed I. C. Engine on CNG, Methane Enriched Biogas and Biogas
,”
Appl. Energy
,
88
(
11
), pp.
3969
3977
.
11.
Bora
,
B. J.
, and
Saha
,
U. K.
,
2016
, “
Optimisation of Injection Timing and Compression Ratio of a Raw Biogas Powered Dual Fuel Diesel Engine
,”
Appl. Therm. Eng.
,
92
, pp.
111
121
.
12.
Bedoya
,
I. D.
,
Saxena
,
S.
,
Cadavid
,
F. J.
, and
Dibble
,
R. W.
,
2013
, “
Numerical Analysis of Biogas Composition Effects on Combustion Parameters and Emissions in Biogas Fueled HCCI Engines for Power Generation
,”
ASME J. Eng. Gas Turbines Power
,
135
(
7
), p.
071503
.
13.
Hotta
,
S. K.
,
Sahoo
,
N.
, and
Mohanty
,
K.
,
2018
, “
Comparative Assessment of a Spark Ignition Engine Fueled With Gasoline and Raw Biogas
,”
Renew. Energy
,
134
, pp.
1
13
.
14.
Youse
,
A.
,
Guo
,
H.
, and
Birouk
,
M.
,
2019
, “
Effect of Diesel Injection Timing on the Combustion of Natural Gas/Diesel Dual-Fuel Engine at Low-High Load and Low-High Speed Conditions
,”
Fuel
,
235
, pp.
838
846
.
15.
Jagadish
,
C.
, and
Gumtapure
,
V.
,
2020
, “
Experimental Studies on Cyclic Variations in a Single Cylinder Diesel Engine Fuelled With Raw Biogas by Dual Mode of Operation
,”
Fuel
,
266
, p.
117062
.
16.
Jagadish
,
C.
, and
Gumtapure
,
V.
,
2019
, “
Environmental Effects Experimental Investigation of Methane-Enriched Biogas in a Single Cylinder Diesel Engine by the Dual Fuel Mode
,”
Energy Sources, Part A
,
14
, pp.
1
14
.
17.
Bora
,
B. J.
,
Saha
,
U. K.
,
Chatterjee
,
S.
, and
Veer
,
V.
,
2014
, “
Effect of Compression Ratio on Performance, Combustion and Emission Characteristics of a Dual Fuel Diesel Engine Run on Raw Biogas
,”
Energy Convers. Manag.
,
87
, pp.
1000
1009
.
18.
Kan
,
X.
,
Zhou
,
D.
,
Yang
,
W.
,
Zhai
,
X.
, and
Wang
,
C. H.
,
2018
, “
An Investigation on Utilization of Biogas and Syngas Produced From Biomass Waste in Premixed Spark Ignition Engine
,”
Appl. Energy
,
212
, pp.
210
222
.
19.
Deheri
,
C.
,
Acharya
,
S. K.
,
Thatoi
,
D. N.
, and
Mohanty
,
A. P.
,
2020
, “
Review Article A Review on Performance of Biogas and Hydrogen on Diesel Engine in Dual Fuel Mode
,”
Fuel
,
260
, p.
116337
.
20.
Cho
,
J.
,
Park
,
S.
, and
Song
,
S.
,
2019
, “
The Effects of the Air-Fuel Ratio on a Stationary Diesel Engine Under Dual-Fuel Conditions and Multi-Objective Optimization
,”
Energy
,
187
, p.
115884
.
21.
Nabi
,
M. N.
,
Rahman
,
S. M. A.
,
Bodisco
,
T. A.
,
Rasul
,
M. G.
,
Ristovski
,
Z. D.
, and
Brown
,
R. J.
,
2019
, “
Assessment of the Use of a Novel Series of Oxygenated Fuels for a Turbocharged Diesel Engine
,”
J. Clean. Prod.
,
217
, pp.
549
558
.
22.
Gopal
,
K.
,
Sathiyagnanam
,
A. P.
,
Rajesh Kumar
,
B.
,
Saravanan
,
S.
,
Rana
,
D.
, and
Sethuramasamyraja
,
B.
,
2018
, “
Prediction of Emissions and Performance of a Diesel Engine Fueled With N-Octanol/Diesel Blends Using Response Surface Methodology
,”
J. Clean. Prod.
,
184
, pp.
423
439
.
23.
Qi
,
D.
,
Leick
,
M.
,
Liu
,
Y.
, and
Lee
,
C. F.
,
2011
, “
Effect of EGR and Injection Timing on Combustion and Emission Characteristics of Split Injection Strategy DI-Diesel Engine Fueled With Biodiesel
,”
Fuel
,
90
(
5
), pp.
1884
1891
.
24.
Caresana
,
F.
,
2011
, “
Impact of Biodiesel Bulk Modulus on Injection Pressure and Injection Timing. The Effect of Residual Pressure
,”
Fuel
,
90
(
2
), pp.
477
485
.
25.
Kumar
,
M. H.
,
Raju
,
V. D.
,
Kishore
,
P. S.
, and
Venu
,
H.
,
2018
, “
Influence of Injection Timing on the Performance, Combustion and Emission Characteristics of Diesel Engine Powered With Tamarind Seed Biodiesel Blend
,”
Int. J. Ambient Energy
,
41
, pp.
1
9
.
26.
Sayin
,
C.
, and
Canakci
,
M.
,
2009
, “
Effects of Injection Timing on the Engine Performance and Exhaust Emissions of a Dual-Fuel Diesel Engine
,”
Energy Convers. Manag.
,
50
(
1
), pp.
203
213
.
27.
Karthic
,
S. V.
,
Kumar
,
M. S.
,
Nataraj
,
G.
, and
Pradeep
,
P.
,
2020
, “
An Assessment on Injection Pressure and Timing to Reduce Emissions on Diesel Engine Powered by Renewable Fuel
,”
J. Clean. Prod.
,
255
, p.
120186
.
28.
Lounici
,
M. S.
,
2018
, “
Experimental Investigation on the Performance and Exhaust Emission of Biogas-Diesel Dual-Fuel Combustion in a CI Engine
.” SAE Technical Paper No. 2689.
29.
Kannan
,
G. R.
, and
Anand
,
R.
,
2011
, “
Experimental Evaluation of DI Diesel Engine Operating With Diestrol at Varying Injection Pressure and Injection Timing
,”
Fuel Process. Technol.
,
92
(
12
), pp.
2252
2263
.
30.
Ambarita
,
H.
,
2017
, “
Performance and Emission Characteristics of a Small Diesel Engine Run in Dual-Fuel (Diesel-Biogas) Mode,” Case Stud
,”
Therm. Eng.
,
10
, pp.
179
191
.
31.
Bora
,
B. J.
, and
Saha
,
U. K.
,
2016
, “
Experimental Evaluation of a Rice Bran Biodiesel-Biogas run Dual Fuel Diesel Engine at Varying Compression Ratios
,”
Renew. Energy
,
87
, pp.
782
790
.
32.
Chandekar
,
A. C.
, and
Debnath
,
B. K.
,
2018
, “
Computational Investigation of Air-Biogas Mixing Device for Different Biogas Substitutions and Engine Load Variations
,”
Renew. Energy
,
127
(
X
), pp.
811
824
.
33.
Bedoya
,
I. D.
,
Arrieta
,
A. A.
, and
Cadavid
,
F. J.
,
2009
, “
Bioresource Technology Effects of Mixing System and Pilot Fuel Quality on Diesel—Biogas Dual Fuel Engine Performance
,”
Bioresour. Technol.
,
100
(
24
), pp.
6624
6629
.
34.
Heywood
,
J. B.
,
2018
,
Internal Combustion Engine Fundamentals
. vol.
21
.
35.
Verma
,
S.
,
Das
,
L. M.
, and
Kaushik
,
S. C.
,
2017
, “
Effects of Varying Composition of Biogas on Performance and Emission Characteristics of Compression Ignition Engine Using Exergy Analysis
,”
Energy Convers. Manag.
,
138
, pp.
346
359
.
36.
Sayin
,
C.
,
Gumus
,
M.
, and
Canakci
,
M.
,
2010
, “
Effect of Fuel Injection Timing on the Emissions of a Direct-Injection (DI) Diesel Engine Fueled With Canola Oil Methyl Ester—Diesel Fuel Blends
,”
Energy Fuels
,
45
(
12
), pp.
2675
2682
.
37.
Rahman
,
K. A.
, and
Ramesh
,
A.
,
2019
, “
Studies on the Effects of Methane Fraction and Injection Strategies in a Biogas Diesel Common Rail Dual Fuel Engine
,”
Fuel
,
236
, pp.
147
165
.