A Study on Combustion Parameters and Exhaust Characteristics in a Diesel Engine Using Alternative Fuels at Different SOI and GPP

Abstract

To encourage the use of alternative fuels while limiting the use of fossil fuels, researchers have focused on using more environmentally friendly fuels. Furthermore, the goal is to improve engine performance to increase energy efficiency. A four-stroke, single-cylinder, diesel engine with a common rail fuel injection system runs with diesel, biodiesel, and biodiesel–alcohol fuel blends. The tests were performed using a constant engine speed of 2000 rpm and three different gas pedal positions (20%, 50% and 80%). It was found that maximum cylinder gas pressure increased in all test fuels with increased gas pedal position (GPP) and advanced injection start time. In general, the maximum heat release rate increased in blended fuels compared to diesel fuel. In addition, it was seen that advanced injection timings caused an increase in ignition delay in all fuel types. In the same test conditions, it was observed that biodiesel–alcohol fuel blends caused an increase in ignition delay by more than 10% compared to diesel fuel (D100), while shortening combustion duration (CD) by more than 10%. A decreasing trend in CO and HC emissions was observed in the use of biodiesel fuel compared to diesel fuel. With the use of biodiesel–alcohol fuel blends, CO₂ emissions tend to decrease. Advanced injection timings caused high NO emissions.

1. Introduction

Despite their widespread application in transportation, agriculture, rail, and marine industries owing to their high efficiency, strong torque, and powerful output, internal combustion diesel engines significantly contribute to harmful emissions because they primarily depend on fossil fuels for energy [1]. Diesel engines operate with high compression ratios and excess air coefficients. Due to the high temperatures reached in the cylinder under these operating conditions, NO_x (nitrogen oxide) gas—which causes acid rain—and PM (particulate matter)—which is produced by the heterogeneous mixture of fuel directly injected into the cylinder—cause emissions. In addition, studies in the literature have stated that CO (carbon monoxide) and HC (hydrocarbon) emissions originating from diesel engines are also released into the environment at certain rates [2,3,4,5]. In this context, researchers are intensively working on strategies such as renewable and environmentally friendly energy sources (biodiesel and alcohol fuels, etc.) [6,7,8,9] and injection parameters [10,11,12,13,14].

In internal combustion engines, sending fuel into the cylinder at the right time, in the right amount, and at the right pressure directly affects the combustion in the cylinder. For this reason, by controlling the combustion in the cylinder, engine performance can be increased, combustion noise can be controlled, and exhaust emissions and fuel consumption can be reduced [15,16]. By optimizing the fuel injection timing and providing sufficient time for the air–fuel mixture to be formed, better combustion and the maximum cylinder gas pressure can be achieved close to Top Dead Center (TDC) [17,18,19].

Biodiesel is considered a promising renewable alternative to petroleum-based diesel, as it can be synthesized from a wide range of vegetable oils and animal fats. Biodiesel fuel can be used directly or by mixing it with diesel fuel at certain rates without any modification in diesel engines. Important features that increase the usage potential of biodiesel fuel include high cetane number, low flash point, non-toxicity, and being a good lubricant. In addition, providing job opportunities by developing local agriculture, contributing to the reduction in external dependency on energy, and contributing to the reduction in exhaust emissions that are harmful to human health and the environment as a result of combustion compared to fossil-based fuel can be listed as important factors for the use of biodiesel. However, the disadvantages of biodiesel fuel in internal combustion diesel engines include its viscosity being twice as high and its calorific value being approximately 20% lower compared to fossil-based diesel fuel [20,21,22,23,24].

Ethanol is considered a light alcohol fuel with high potential for use in internal combustion engines due to its ability to be produced from renewable energy sources and having properties similar to fossil fuels. Ethanol can improve combustion in the cylinder and increase brake thermal efficiency thanks to its high oxygen content and low viscosity compared to fossil diesel fuel. And also, the high latent heat of vaporization of ethanol causes an increase in volumetric efficiency thanks to its cooling effect in the cylinder. On the other hand, due to the OH (hydroxyl) molecules in ethanol, mixing it with fossil diesel fuel is limited to a certain volume. As a result of mixing high amounts of ethanol with fossil diesel fuel, phase separation occurs. This is primarily due to the more polar nature of ethanol and the lower polarity of diesel fuel. Therefore, the miscibility of ethanol with a nonpolar liquid like diesel is limited. For this reason, the fuels separate from each other into two different layers. In the studies conducted in the literature, it is recommended to add solvent to ethanol–diesel fuel mixtures to prevent phase separation.

On the other hand, researchers are intensively working on the use of isopropanol and butanol, which have higher calorific value and cetane number than ethanol, in internal combustion engines. Since isopropanol and butanol form a more stable mixture with fossil-based diesel fuel, they have a higher potential for addition to fossil-based fuels. In addition, isopropanol and butanol fuels can be produced from biomass or crude oil. It is stated that isopropanol and butanol are safer to store and transport because their flash points are higher than those of ethanol. On the other hand, the fact that the production costs of isopropanol and butanol are higher than those of ethanol is seen as a disadvantage [25,26,27,28,29,30,31].

In the literature, Koçyiğit S. et al. [32] studied the effects of using diesel, diesel–ethanol, and diesel–ethanol–propolis mixtures on combustion and exhaust emissions at a 3000 rpm constant engine speed and under different engine loads in a compression ignition engine. They reported that the use of blended fuels resulted in a decrease in heat release rate (HRR) and cylinder gas pressure (CGP) compared to diesel fuels. Blended fuels resulted in a decrease in CO emissions but an increase in CO₂ emissions. They also noted that brake specific fuel consumption (BSFC) decreased for the test fuels with increasing engine load. Ferreira V. P. et al. [33], the effects of using diesel–biodiesel–ethanol fuel blends in a four-stroke, single-cylinder diesel engine on combustion and exhaust emissions at constant engine speed. The mixed fuels caused an increase in HC and CO; however, a decrease in CO₂ emissions when compared to diesel fuel. It was stated that NO_x decreased as a result of the experiments where the ethanol ratio in blended fuels was lower than 10%, but NO_x increased at points higher than 10%. Recently, Sayyed S. et al. [34] investigated the effects of diesel–biodiesel–butanol fuel mixtures in a single-cylinder diesel engine on engine performance, combustion, and exhaust emission characteristics at constant engine speed and different engine loads. The use of blended fuels in the test conditions caused a decrease in thermal efficiency, but significantly increased the exhaust gas temperature (EGT). The blended fuels caused a decrease in HC and CO emissions, while increasing NO_x and CO₂ emissions, compared to diesel fuel in all test conditions.

In another study, Li J. et al. [35] investigated the effects of diesel–ethanol–butanol fuel mixtures at different injection timings on engine performance, combustion, and exhaust emission characteristics in a four-stroke, four-cylinder diesel engine. The cylinder gas pressure, HRR, and cylinder temperature increased with the advancement in injection timing for each fuel. They stated that the increase in alcohol fuel ratio in the blended fuels increased specific fuel consumption and thermal efficiency compared to diesel fuel. They found that at the same injection start timing, the decrease in diesel ratio in the blended fuels reduced NO_x emissions. Liang J. et al. [36] investigated the effects of injection timing on engine performance, combustion, and exhaust emission characteristics using ethanol–diesel fuel blends in a single-cylinder compression ignition engine. The use of blended fuel increased the ignition delay (ID) but shortened the combustion period. They stated that delaying the injection timing reduced the thermal efficiency in all fuels. The blended fuels were higher than diesel fuel in NO_x emissions at the same injection start timing. Murcak A. et al. [37], in their study, investigated the effects of using diesel–ethanol fuel mixtures in a single cylinder diesel engine on engine performance at different injection times, different engine speeds, and full engine load. The use of fuel mixtures in the conditions tested caused an increase in BSFC, but there was an increase in engine power with the use of a certain ratio of ethanol and diesel fuel.

Park H. S. et al. [38] investigated the effects of injection timing on combustion characteristics in a four-stroke, four-cylinder diesel engine using diesel–ethanol blend fuels. They observed that the CGP obtained under the same test conditions were close to each other, but for all fuel types, advancing the injection timing improved the CGP and HRR. Jain A. et al. [39], in their study, investigated the effects of using diesel–biodiesel–alcohol fuel blends at different injection start times on engine performance, combustion, and exhaust emission in a compression ignition engine. They reported that brake thermal efficiency, HRR, and ID decreased with the use of blended fuel compared to diesel fuel, but they found an increase in CGP. The early injection timing increased HC and CO emissions for the same fuel type. Karthic S. V. et al. [40] observed the impact of biodiesel–diesel fuel blends on engine performance, combustion, and exhaust emission characteristics at different start of injection (SOI) and injection pressures in a single cylinder diesel engine. They stated that diesel fuel has higher thermal efficiency compared to blended fuels under the same injection conditions and that specific fuel consumption values are lower. They found that early fuel injection timing improves HRR in all fuel types. The early fuel injection timing causes an increase in NO emissions at the same injection pressure. Li G. et al. [41] researched the effects of different SOI on engine performance, combustion, and exhaust emissions when using diesel–ethanol–propanol–butanol fuel mixtures in a single-cylinder diesel engine. It was reported that early SOI improved CGP and HRR in all fuel types. Especially adding 15% alcohol (ethanol–propanol–butanol) to diesel fuel improved thermal efficiency and reduced BSFC. HC emissions decreased with early SOI.

As mentioned above, experiments have been conducted in the literature using diesel–biodiesel–alcohol fuel mixtures in different ratios and under different engine conditions. However, it was observed that the investigation of the effect of injection timing on biodiesel and biodiesel–alcohol blends under the same conditions was incomplete. Therefore, this study focuses on the injection parameter, which is considered to be one of the parameters that has a significant impact on biodiesel usage. In addition, biodiesel fuel has been used in higher mixture ratios compared to alcohol fuels. As understood from the studies conducted in the literature, when low-carbon alcohol fuels are mixed with fossil-based diesel fuel, the formation of phase separation has limited the mixture ratio. In addition, the use of biodiesel fuel alone or in high mixture ratios increases the viscosity of the fuel to be injected into the cylinder and makes evaporation in the cylinder difficult. For this reason, biodiesel fuel produced from waste cooking oil, which was selected as an alternative fuel, was mixed with alcohol fuel in order to obtain suitable fuel properties for use in internal combustion engines. In addition, it was aimed to examine the effects by calibrating the injection parameters. In this context, five different fuel types were selected in the study, taking diesel fuel as reference, namely 100% biodiesel, 80% biodiesel + 20% ethanol, 80% biodiesel + 20% isopropanol, and 80% biodiesel + 20% butanol fuels. The aim was to obtain fuels with properties close to D100 using these fuel types. In addition, the five selected fuel types were examined separately, starting from standard operating conditions, at different injection start times (±5 °CA), and at three different gas pedal positions (20%, 50% and 80%). The experiments were conducted by holding the amount of fuel delivered to the cylinder, coolant temperature, and oil temperature constant to directly observe the effects of fuel changes and injection timing. The data obtained as a result of the experiments were compared with standard fuel and standard operating parameters.

2. Experimental Procedure

2.1. Experimental Fuels

In this study, fossil-based diesel fuel, which was accepted as the reference fuel to be used in engine tests, was supplied by Shell in Turkey, an international company. The biodiesel fuel produced from waste cooking oil used in the experiments was supplied by Deha Biodiesel, a national fuel producer. In addition, three alcohol fuels (ethanol, isopropanol, and butanol) were purchased to be mixed into the biodiesel fuel. Ethanol, which has a purity of more than 99.8%, was purchased from the Company Honeywell. In addition, isopropanol, which has a purity of more than 99.9%, was supplied by Ekiciler Chemical, while butanol fuel was purchased from Company Sigma-Aldrich, and its purity level is higher than 99.9%.

A total of five different fuels were used in the experiments. The fuels used are given below.

• 100% fossil fuel,
• 100% biodiesel (produced from waste cooking oil),
• 80% biodiesel + 20% ethanol,
• 80% biodiesel + 20% isopropanol,
• 80% biodiesel + 20% butanol.

The fuels used in the study were named based on the fuels they contain by volume. The fuel containing 100% fossil-based diesel by volume was named D100, while the fuel containing 100% biodiesel by volume was named BD100. The fuel mixture containing 20% ethanol and 80% biodiesel by volume was accepted as BD80E20, while the fuel mixture consisting of 20% isopropanol and 80% biodiesel by volume was named BD80P20. In addition, the fuel prepared by mixing 20% butanol and 80% biodiesel by volume was named BD80B20. In this study, beakers were used to prepare fuel mixtures. Large beakers were used due to the high proportion of biodiesel fuel in the mixture, while smaller beakers were used due to the lower proportion of alcohol in the mixture. The fuels were mixed and stored after measuring the determined volumetric values in the beakers. The prepared fuels are given in Figure 1.

The prepared fuel was stored under conditions that prevented its exposure to air. The fuels were mixed until homogeneity was achieved in the mixed fuels, and then observations were made for phase separation. Throughout the 10-day observation period, no distinct phase separation was detected in the test fuels. Figure 2 illustrates the appearance of the fuels after 240 h.

The characteristics of the test fuels prepared for the study were analyzed at the TÜBİTAK Marmara Research Center (MAM) laboratories in Turkey in compliance with international testing standards. The fuel specifications are presented in

Table 1. Properties of test fuels.

	D100	BD100	BD80E20	BD80P20	BD80B20
Chemical Formula	C12H26	C18.75H36O2	C8.9H18.37O1.41	C10.56 H21.44 O1.48	C10.78H21.95O1.45
Density (kg/m3)	831.7	884.2	864.7	864.4	869.5
Higher calorific value (MJ/kg)	46.89	39.79	37.32	38.19	38.86
Lower calorific value (MJ/kg)	43.76	38.65	36.87	37.5	37.98
Flash point (°C)	63.5	163.5	20	20	43.5
Cetane number	54.9	53.4	47.5	50.1	50.5
Kinematic viscosity (mm2/s)	3.25	4.37	2.88	3.15	3.33
C (%) mass	84.7	76.79	72.29	73.74	74.12
H (%) mass	15.3	12.28	12.43	12.47	12.57
O (%) mass	-	10.93	15.28	13.79	13.31

. Additionally, some properties of the alcohols used are given in

Table 2. Some properties of alcohol fuels [42,43,44].

	Ethanol	Isopropanol	Butanol
Chemical Formula	C2H5OH	C3H7OH	C4H9OH
Density (kg/m3)	789	787	840
Lower calorific value (MJ/kg)	27	30.4	33.6
Cetane number	8	-	17
Kinematic viscosity (mm2/s)	1.5	1.69	3.6
Flash point (°C)	-	12	35
C (%) mass	52.2	60	64.9
H (%) mass	13.1	13.4	13.5
O (%) mass	34.7	26.6	21.6

.

2.2. Experimental System

The engine tests for this study were carried out in the engine test laboratory located at the TUBITAK Rail Transportation Technologies Institute in Turkey. The main image of the experimental setup is given in Figure 3, and a schematic view of the engine setup is given in Figure 4.

In the experimental system, the intake air temperature conditioner, intake air pressure conditioner, and lubrication oil conditioner systems can also operate in a fully integrated manner and can be controlled via electronic control unit (ECU) maps. Through the ECU’s configurable maps, parameters such as the start of injection, injection quantity, and injection pressure can be adjusted in real time, allowing the effects of these variables on engine performance to be monitored using the test engine. In addition, torque is measured using the HBM T40B torque flange mounted between the engine and the dynamometer shaft during the experiments. The devices used in the experimental system and their accuracy values are given in

Table 3. Devices used in the experiments and their accuracy values.

Measurement	Device	Accuracy
Torque	AVL-HBM Torque Flange	±0.1%
Engine Speed	AVL Encoder	≤±0.1 °CA
Test Cell Humidity and Temperature	Vaisala-HMT 330	±1% RH, ±0.2 °C
In-cylinder Pressure	AVL-Xion	0.05 °CA
Injection timing	Angle Encoder	±0.1 °CA
Fuel Consumption	AVL-735	<0.15%
Temperature Sensors	PT100 (K Type)	≤±1%

.

An AVL eddy-current dynamometer was employed in the engine test setup. This dynamometer is capable of measuring engine torque within the range of 0 to 190 Nm and can operate at engine speeds ranging from 0 to 8000 rpm. The engine dynamometer was controlled using the AVL PUMA automation system.

In the experiments, the GH14D type piezoelectric sensor with an accuracy of ±0.01% up to 250 bar was integrated into the glow plug hole of a single-cylinder research engine to measure the cylinder gas pressure, and pressure measurement was carried out. In addition, the injection current can be measured with a Fluke-600 brand Amperemeter with an accuracy of ±0.3% up to 100 A placed on the injector cable, and the opening and closing timings of the injector can be observed. With the observed injector current, the injection start timings, the end of the injection, and the total injection amount were checked. The measured pressure and current changes are read using AVL Indicom.

Engine performance evaluations were carried out on a John Deere single-cylinder test engine featuring a four-stroke cycle, water-cooling system, and common rail fuel injection. The engine has a displacement of 1.205 L and a three-valve configuration per cylinder (two intake valves and one exhaust valve). Technical parameters of the engine are summarized in

Table 4. Specifications of the test engine.

Parameter	Specification
Brand	John Deere
Engine Type	Four-Stroke
Number of Cylinders	Single Cylinder
Cooling System	Water-Cooled
Fuel System	Common Rail Direct Injection—1800 bar
Cylinder Volume	1205 cm3
Valves	3 (2 intake–1 exhaust)—(OHV)
Max. Cylinder Pressure	190 bar
Max. Engine Speed	2500 rpm
Max. Power	50 kW
Max. Torque	160 Nm
Bore	106.5 mm
Stroke	127 mm
Compression Ratio	16.14

.

In the experiments, the AVL DITEST MDS-650 device, which can operate in environments with a temperature of 40 °C and 90% humidity for emission measurement, was used. CO, CO₂, HC, and NO emissions were measured with the emission device used. The measured exhaust gas values were measured in units of particles per million (ppm). The accuracy values of the AVL DITEST exhaust emission device are given in

Table 5. Exhaust emission device accuracies.

Exhaust Emissions	Measuring Device	Range	Resolution	Accuracy Value
CO	AVL DITEST	0–15%	0.01% vol.	≤±3%
CO2	AVL DITEST	0–20%	0.01% vol.	≤±3%
NO	AVL DITEST	0–5000 ppm	1 ppm vol.	≤±1%
HC	AVL DITEST	0–30,000 ppm	1 ppm vol.	≤±3%

.

2.3. Conducting Experiments

The experiments within the scope of this study were carried out in accordance with the test environment specified in ISO 8178 within TUBITAK RUTE and the power measurement method reported in ISO 14396.

Before starting the experiments, calibration checks were made for the exhaust emission device. Then, the engine was operated for a certain period of time with D100 reference fuel until it reached a stable state. During the engine tests, cylinder gas pressure data were recorded at 7200 points per cycle, with measurements taken every 0.1 degrees of crank angle (°CA). For each test condition, experiments were conducted in triplicate. In every experiment, data were collected over 100 cycles, and the averaged values from these cycles were calculated and utilized in the study. In the engine tests, the cooling water temperature, fuel temperature, and oil temperature values were kept constant throughout each experiment at 80 °C, 25 °C, and 80 °C, respectively.

The experiments were carried out by changing the fuel injection start timing at approximately partial engine loads (20%, 50% and 80% gas pedal positions) while the engine speed was kept constant at 2000 rpm. Considering the general operating conditions of diesel engines in the selection of GPP in the experimental conditions, 20% was preferred as the lowest partial load, 80% as the highest partial load, and 50% as the medium partial load. D100, BD100, BD80E20, BD80P20, and BD80B20 were used in the experiments, respectively. In order to observe the effect of the SOI, the fuel was first injected before top dead center (TDC) at the standard SOI, and experiments were carried out. Then, in order to determine the effect of the fuel injection timing on the engine, the timing was changed by ±5 °CA, and the data were obtained. The injection start times are given in

Table 6. Injection start times.

Engine Speed (rpm)	Gas Pedal Position (%)	5 °CA Retarded	Standard Injection Start Time	5 °CA Advanced
2000	20	9.8	14.8	19.8
	50	10.4	15.4	20.4
	80	11.9	16.9	21.9

.

2.4. Calculation Methods Used in Experiments

Calculations of data obtained during engine tests were made with fully integrated computer control. Formulas used in calculations are given below.

2.4.1. Calculation of HRR

The heat released within the cylinder at each crank angle can be quantified based on the in-cylinder pressure. The calculation of the heat release rate was performed with reference to the first law of thermodynamics.

$${\displaystyle \frac{d\mathrm{Q}}{\mathrm{d}\mathrm{t}}}={\displaystyle \frac{d\mathrm{W}}{\mathrm{d}\mathrm{t}}}+{\displaystyle \frac{\mathrm{d}\mathrm{U}}{\mathrm{d}\mathrm{t}}}$$

(1)

The expressions in Equation (1) are explained as follows:

${\displaystyle \frac{d\mathrm{Q}}{\mathrm{d}\mathrm{t}}}$: Net in-cylinder heat transfer (kJ/s),

${\displaystyle \frac{d\mathrm{W}}{\mathrm{d}\mathrm{t}}}$: Work attained by the piston in the cylinder (kJ/s),

${\displaystyle \frac{\mathrm{d}\mathrm{U}}{\mathrm{d}\mathrm{t}}}$: Change in internal energy (kJ/s),

In Equation (1), mass changes are neglected, and the expressions can be written in the following formulation:

$${\displaystyle \frac{d\mathrm{W}}{\mathrm{d}\mathrm{t}}}=\mathrm{P}.{\displaystyle \frac{\mathrm{d}\mathrm{V}}{\mathrm{d}\mathrm{t}}}$$

(2)

$${\displaystyle \frac{\mathrm{d}\mathrm{U}}{\mathrm{d}\mathrm{t}}}=\mathrm{m}.{\mathrm{C}}_{\mathrm{v}}.{\displaystyle \frac{\mathrm{d}\mathrm{T}}{\mathrm{d}\mathrm{t}}}$$

(3)

If the expression of the change with respect to time ($\mathrm{d}\mathrm{t}$), in Equation (3), is expressed in terms of crankshaft ($\mathrm{d}\mathsf{\theta }$), Equation (4) is obtained.

$${\displaystyle \frac{\mathrm{d}\mathrm{U}}{\mathrm{d}\mathsf{\theta }}}=m.C_v.{\displaystyle \frac{\mathrm{d}\mathrm{T}}{\mathrm{d}\mathsf{\theta }}}$$

(4)

$${\displaystyle \frac{\mathrm{d}\mathrm{T}}{\mathrm{d}\mathsf{\theta }}}={\displaystyle \frac{d\mathrm{W}}{\mathrm{m}.\mathrm{R}}}$$

(5)

$$R=C_p-C_v$$

(6)

$$k={\displaystyle \frac{C_p}{C_v}}$$

(7)

Equation (8) is obtained by performing the necessary operations in Equations (6) and (7).

$$k-1={\displaystyle \frac{R}{C_v}}$$

(8)

By using Equations (2), (4) and (8) in Equation (1), Equation (9) can be obtained to be used in heat distribution calculations for the first law.

$$Q={\displaystyle \frac{\mathrm{k}}{\mathrm{k}-1}}.\mathrm{P}.{\displaystyle \frac{\mathrm{d}\mathrm{V}}{\mathrm{d}\mathsf{\theta }}}+{\displaystyle \frac{1}{\mathrm{k}-1}}.\mathrm{V}.{\displaystyle \frac{\mathrm{d}\mathrm{P}}{\mathrm{d}\mathsf{\theta }}}$$

(9)

The expressions of the formulas are explained as follows:

$\mathrm{P}$: Cylinder gas pressure (Pa),

$V$: Cylinder volume (m³),

$T$: Cylinder gas temperature (K),

$\theta$: Crank angle (°CA),

$C_v$: Specific heat at constant volume (kJ/kgK),

$C_p$: Specific heat at constant pressure (kJ/kgK),

$m$: Mass (kg),

$R$: Gas constant (kJ/kgK),

$k$: Specific heat ratio,

2.4.2. Calculation of Combustion Stages

The combustion event in diesel engines initiates with the injection of fuel into the cylinder and proceeds until the HRR decreases to near zero. This process relies on the high-pressure injection of fuel into the compressed air, which reaches increased temperature and pressure during the compression stroke.

In the studies conducted in the literature, different approaches and interpretations were made for the start of combustion in the cylinder. In the evaluations made, the start of fuel injection in the cylinder can be accepted as the moment when the injector needle is lifted. However, since it is not practical to make this determination, it is also accepted as the moment when current is sent to the injector. In this study, the point where the injector is energized is accepted as the start of injection. In addition, the point where the injector energization ends is accepted as the end of injection. Equation (10) was used in the calculation of the injection period.

$$FID={\theta }_{PS}-{\theta }_{PB}$$

(10)

where the expressions are explained as follows:

$FID$: Fuel injection duration (°CA),

${\theta }_{PB}$: Start of fuel injection (°CA),

${\theta }_{PS}$: End of fuel injection (°CA),

When examining methods to determine the start of combustion, one approach defines it as the point where the heat release rate curve rises from its minimum to a positive value. Alternatively, some consider the start of combustion as the moment when 5% or 10% of the total fuel, as calculated from the HRR, has been burnt. Ignition delay (ID) can be described as the interval between the fuel injection event—during which the ID process occurs—and the instant the fuel first ignites in the cylinder. Equation (11) was utilized to calculate the ignition delay.

$$ID={\theta }_{YB}-{\theta }_{SOI}$$

(11)

where the expressions are explained as follows:

$ID$: Ignition delay (°CA),

${\theta }_{YB}$: Start of combustion as crankshaft angle (°CA),

${\theta }_{YB}$: Start of combustion injection as crankshaft angle (°CA),

The total combustion duration in the cylinder is defined as the time interval from the start of combustion until its completion within the cylinder. In the literature, the end of the combustion event is commonly defined as the point at which 90% to 95% of the fuel has been consumed. In other words, the combustion termination corresponds to the crank angle where 90% to 95% of the cumulative heat release has occurred in the cylinder [45,46,47]. In this study, the point where the cumulative heat release is 90% was taken as the end of combustion. It was calculated with the expression in Equation (12).

$$CD={\theta }_{YS}-{\theta }_{YB}$$

(12)

where the expressions are explained as follows:

$CD$: Combustion duration (°CA),

${\theta }_{YS}$: End of combustion as crankshaft angle (°CA),

${\theta }_{YB}$: Start of combustion as crankshaft angle (°CA),

Definitions regarding the combustion process are given in Figure 5.

Within the scope of this study, ID, start of combustion, end of combustion, and CD calculations were made by the system using the data obtained from the crank angle in the AVL Concerto 5R1.6 software.

2.4.3. Emission Calculation

In the experiments, CO, CO₂, NO, and HC emissions were measured using the AVL DITEST device. The measurements made with the emission device in the experiments will be made in ppm and % units. However, for comparison with the studies in the literature and for a better interpretation of the experimental results, the units of exhaust emissions will be converted to g/kWh. The following model will be used for unit conversion.

$$C_x{H}_y{O}_z+Q_{min}\lambda \left(O_2+3.76N_2\right)\to a{CO}_2+b{H}_{2}O+c{N}_2+d{O}_2$$

(13)

$$a=x$$

(14)

$$b=y/2$$

(15)

$$c=Q_{min}.\lambda .3.76$$

(16)

$$d={\displaystyle \frac{\left[\left(z+Q_{min}.\lambda .2\right)-2a-b\right]}{2}}$$

(17)

$$n_T=a+b+c+d$$

(18)

$${CO}_{2}mole fraction={\displaystyle \frac{n_{{CO}_2}}{n_T}}$$

(19)

$$O_{2}mole fraction={\displaystyle \frac{n_{O_2}}{n_T}}$$

(20)

$$N_{2}mole fraction={\displaystyle \frac{n_{N_2}}{n_T}}$$

(21)

$$H_{2}O mole fraction={\displaystyle \frac{n_{H_{2}O}}{n_T}}$$

(22)

$$m_{{CO}_2}={CO}_2 mole fraction \ast M_{{aCO}_2}$$

(23)

$$m_{O_2}=O_2 mole fraction \ast M_{{aO}_2}$$

(24)

$$m_{N_2}=N_2 mole fraction \ast M_{{aN}_2}$$

(25)

$$m_{H_{2}O}=H_{2}O mole fraction \ast M_{{aH}_{2}O}$$

(26)

$$M_{a_{exhaust}}=m_{{CO}_2}+m_{O_2}+m_{N_2}+m_{H_{2}O}$$

(27)

$$n_{exhaust}={\displaystyle \frac{M_{exhaust}}{M_{a_{exhaust}}}}$$

(28)

$$HC({\displaystyle \frac{g}{kWh}})={\displaystyle \frac{n_{exhaust}\ast HC\left(ppm\right)\ast M_{aHC}}{Pe\ast 1000000}}$$

(29)

$${CO}_2({\displaystyle \frac{g}{kWh}})={\displaystyle \frac{n_{exhaust}\ast {CO}_2\left(ppm\right)\ast M_{{aCO}_2}}{Pe\ast 1000000}}$$

(30)

$$CO({\displaystyle \frac{g}{kWh}})={\displaystyle \frac{n_{exhaust}\ast CO\left(ppm\right)\ast M_{aCO}}{Pe\ast 1000000}}$$

(31)

$$NO({\displaystyle \frac{g}{kWh}})={\displaystyle \frac{n_{exhaust}\ast NO\left(ppm\right)\ast M_{aNO}}{Pe\ast 1000000}}$$

(32)

where the following applies:

$a$: Number of moles of CO₂ in the combustion equation (mol),

$b$: The number of moles of H₂O in the combustion equation (mol),

$c$: The number of moles of N₂ in the combustion equation (mol),

$d$: The number of moles of O₂ in the combustion equation (mol),

$\lambda$: Air excess coefficient,

${\dot{m}}_{{CO}_2}$: Amount of CO₂ in one mole of exhaust gas (g/h),

${\dot{m}}_{H_{2}O}$: Amount of H₂O in one mole of exhaust gas (g/h),

${\dot{m}}_{N_2}$: The amount of N₂ in one mole of exhaust gas (g/h),

${\dot{m}}_{O_2}$: The amount of O₂ in one mole of exhaust gas (g/h),

$M_{aCO}$: CO molecular weight (g/mol),

$M_{{aCO}_2}$: CO₂ molecular weight (g/mol),

$M_{a_{exhaust}}$: Exhaust gas molecular weight (g/mol),

$M_{aHC}$: HC molecular weight (g/mol),

$M_{{aH}_{2}O}$: H₂O molecular weight (g/mol),

$M_{{aN}_2}$: N₂ molecular weight (g/mol),

$M_{aNO}$: Molecular weight of NO (g/mol),

$M_{{aO}_2}$: O₂ molecular weight (g/mol),

${\dot{M}}_{exhaust}$: Weight of exhaust gas (g/h),

$n_{{CO}_2}$: Number of moles of CO₂ (mol),

${\dot{n}}_{exhaust}$: Number of moles of exhaust gas (mol/h)

$n_{H_{2}O}$: Number of moles of H₂O (mol),

$n_{N_2}$: Number of moles of N₂ (mol),

$n_{O_2}$: Number of moles of O₂ (mol),

$n_T$: Total number of moles of products (mol),

$Pe$: Effective power (kW),

$y$: Number of hydrogens in fuel compound,

$x$: Number of carbons in fuel compound,

$Q_{min}$: The number of moles of oxygen required for complete combustion of the fuel (mol).

2.4.4. Calculating Standard Deviation

The accuracy of the data obtained as a result of the experiments can be directly affected by the calibration status of the devices used and the environmental conditions. In this study, the averages of the data obtained as a result of the engine tests were taken, and the deviations from the average values were determined by calculating the standard deviation (SD) values.

$$AVR={\displaystyle \frac{1}{N}}{\sum }_{x=1}^N{A}_x$$

(33)

$$SD=\sqrt{\left[{\displaystyle \frac{1}{N-1}}{\sum }_{x=1}^N{(A_x-AVR)}^2\right]}$$

(34)

where the expressions are explained as follows:

$SD$: Standard deviation value,

$N$: Number of measurements,

$A$: Measured value,

$AVR$: Average of measured values,

The average values of the parameters measured in the experiments were calculated from the formula in Equation (33) and used in Equation (34), and then the SD values were calculated.

2.5. Cost Analysis

Five different fuel types were used in this study. The per-liter prices of each fuel, calculated using unit prices as reference, are given in

Table 7. Cost analysis.

GPP	SOI	Fuel Types	Cost ($/Liter)	Cost ($/kWh)
20%	Retarded	D100	1.26	0.73
		BD100	1.07	0.52
		BD80E20	2.48	1.76
		BD80P20	1.58	1.17
		BD80B20	4.34	3.31
	Standard	D100	1.26	0.72
		BD100	1.07	0.49
		BD80E20	2.48	1.68
		BD80P20	1.58	1.07
		BD80B20	4.34	3.04
	Advanced	D100	1.26	0.67
		BD100	1.07	0.47
		BD80E20	2.48	1.64
		BD80P20	1.58	1.07
		BD80B20	4.34	3.01
50%	Retarded	D100	1.26	0.59
		BD100	1.07	0.47
		BD80E20	2.48	1.29
		BD80P20	1.58	0.84
		BD80B20	4.34	2.32
	Standard	D100	1.26	0.56
		BD100	1.07	0.43
		BD80E20	2.48	1.22
		BD80P20	1.58	0.8
		BD80B20	4.34	2.21
	Advanced	D100	1.26	0.56
		BD100	1.07	0.42
		BD80E20	2.48	1.22
		BD80P20	1.58	0.81
		BD80B20	4.34	2.22
80%	Retarded	D100	1.26	0.56
		BD100	1.07	0.44
		BD80E20	2.48	1.09
		BD80P20	1.58	0.71
		BD80B20	4.34	1.94
	Standard	D100	1.26	0.53
		BD100	1.07	0.42
		BD80E20	2.48	1.05
		BD80P20	1.58	0.67
		BD80B20	4.34	1.87
	Advanced	D100	1.26	0.51
		BD100	1.07	0.41
		BD80E20	2.48	1.03
		BD80P20	1.58	0.67
		BD80B20	4.34	1.84

. While the current prices in [48] for D100 and in [49] for BD100 were taken as reference, the pricing of alcohol fuels was taken from the companies that supplied the fuels.

3. Result and Discussion

3.1. Cylinder Gas Pressure

The cylinder gas pressure resulting from combustion in the cylinder is considered an important indicator for the evaluation of combustion parameters. Figure 6 shows the changes in CGP values at different pedal positions at different SOI timings. The effect of SOI was clearly seen in all fuel types and pedal positions. While it was determined that there was an improvement of more than 15% in maximum CGP values with advanced SOI at the same low pedal position, it was observed that the improvement in maximum CGP approached 50% at high pedal positions. In the experiments conducted at 20% pedal position, the highest maximum CGP was obtained with BD80B20 at 66.8 bar at 19.8 °CA bTDC (before Top Dead Center) and 3.6 °CA aTDC (after Top Dead Center) at the SOI. The increase obtained by advancing the SOI is thought to be due to the start of combustion (SOC) before the end of the compression stroke. In all test conditions, the highest maximum CGP was measured by advancing the SOI 5 °CA at 80% pedal position as 90.1 bar using BD80E20 fuel.

The rise in maximum CGP observed at all injection start timings with increasing GPP is attributed to the higher amount of fuel injected into the cylinder. In addition, as a result of the experiments conducted at the same SOI and gas pedal positions, it was observed that the use of blended fuels increased maximum CGP under certain test conditions compared to D100 fuel. This situation is due to the oxygen molecules contained in the blended fuels improving combustion in the cylinder. In retarded SOI applications, the maximum CGP values were observed to be lower compared to other SOI timings. It can be explained by the shift of SOC towards the expansion stroke, and the maximum CGP values are obtained after 10 °CA from TDC. The maximum CGP value, measured °CA, where CGP obtained (ACGP) and standard deviation (SD) values measured in the experiments are given in

Table 8. CGP, SD, and ACGP values.

GPP	SOI Application	Value	Unit	D100	BD100	BD80E20	BD80P20	BD80B20
20%	Retarded	CGP	bar	49.7	49.8	47.9	44.4	43.4
		SD	-	1	1.1	1.2	1.4	1.6
		ACGP	aTDC	11.6	10.5	12.2	14.9	15.8
	Standard	CGP	bar	59.4	58.9	58.8	59.3	59.6
		SD	-	1.6	1.7	1.6	1.9	1.8
		ACGP	aTDC	6	7.4	6.2	7.3	7.8
	Advanced	CGP	bar	66.5	66.3	64.4	66.1	66.8
		SD	-	2.1	2.1	2.1	3	2.8
		ACGP	aTDC	3.7	3.7	2.3	3.4	3.6
50%	Retarded	CGP	bar	54.6	54.7	54.9	55.7	57.8
		SD	-	1.6	1.3	1.5	1.2	1.6
		ACGP	aTDC	11.4	11.5	11.5	11.4	11.5
	Standard	CGP	bar	64.5	64.5	64.7	67.8	66.4
		SD	-	1.7	2	2.6	2.3	2.8
		ACGP	aTDC	6.7	7	6.8	7.4	7.1
	Advanced	CGP	bar	76.6	74.7	74.9	75.4	76.8
		SD	-	2	2.2	2.4	3.5	2.7
		ACGP	aTDC	4.7	5	3.3	2.9	3.7
80%	Retarded	CGP	bar	60	62.3	61.2	62.2	61.7
		SD	-	1.3	1.4	1.7	2	2
		ACGP	aTDC	10.9	10.7	11.5	11.9	9.5
	Standard	CGP	bar	75.9	75.4	74.7	77.9	76.7
		SD	-	1.8	1.8	2.3	2.6	2.3
		ACGP	aTDC	10.4	9	9.6	7.6	9.6
	Advanced	CGP	bar	87.4	87.6	90.1	87.4	87.4
		SD	-	2.3	2.4	3	2.9	3.1
		ACGP	aTDC	6.1	6.2	4.9	5.2	5.6

.

In addition, when the GPP values were evaluated, it was seen that the maximum CGP increased as the GPP amount increased, as expected. While increasing from 20% GPP to 50% GPP in all fuel types resulted in an approximately 10% improvement in maximum cylinder gas pressures, increasing from 50% GPP to 80% GPP resulted in an approximately 15% increase in maximum cylinder gas pressure. The reason for the increase in maximum cylinder gas pressure due to the increase in GPP is the increase in the amount of fuel sent into the cylinder and the resulting increase in the amount of energy, resulting in higher temperatures and pressures.

The reduction in CGPs observed with delayed SOI during tests involving alcohol blends at low engine loads is attributed to the high heat of vaporization of alcohol. However, this effect diminishes as in-cylinder temperatures rise with increasing engine load.

3.2. Heat Release Rate

HRR represents the amount of heat generated during combustion inside the cylinder as a function of crankshaft angle. Figure 7 illustrates the variation in HRR at different GPP and SOI. Due to the high cetane numbers of D100 and BD100 fuels, combustion initiation occurred earlier compared to other fuels. Consequently, combustion was found to be more controlled, resulting in lower peak HRR values for D100 and BD100 relative to blended fuels.

It was observed that the maximum HRR peaks occurred later when using BD80E20, BD80P20, and BD80B20 fuels compared to D100 and BD100. This delay is attributed to the increased ID and subsequent late combustion start caused by the lower cetane numbers of the alcohol-containing fuels. Furthermore, the highest maximum HRR values recorded under all test conditions at 50% gas pedal position and 20.4 °CA bTDC SOI timing were 202.5 J/°CA and 195.4 J/°CA for BD80P20 and BD80B20 fuels, respectively. The increase in maximum HRR for these fuels is linked to prolonged ID, leading to fuel accumulation and more rapid combustion. Additionally, the reduction in CD observed with blended fuels shortens the total burn time, thereby increasing the peak HRR.

In the experiments conducted in all GPP with standard SOI application, the highest maximum HRR was seen as 178.9 J/°CA at 1 °CA ATDC in BD80P20 fuel, while it was found as 172 J/°CA at 3 °CA ATDC in BD80B20 fuel. While the maximum HRR showed an improvement of more than 10% in 20% GPP in BD80E20 fuel compared to D100 fuel, it was determined that the maximum HRRs were close to each other in 50% and 80% GPP. On the other hand, when the effect of GPP was examined, the lowest maximum HRR in all fuel types was obtained in 80% GPP. This was an expected situation due to the longer combustion duration resulting from the higher total fuel amount sent to the cylinder in 80% GPP. In addition, in all fuel types, heat releases started earlier in 80% GPP compared to the test results conducted in other GPPs due to the earlier fuel injection into the cylinder in standard SOI.

3.3. Ignition Delay

Ignition delay (ID) is defined as the time interval between the start of fuel injection into the cylinder and the instant when the fuel first ignites. Figure 8 illustrates the variation in ID values at different GPP and SOI. ID is influenced primarily by the fuel’s cetane number and latent heat of vaporization. An increase in gas pedal position led to a reduction in ID across all fuel types, likely due to elevated in-cylinder temperatures. ID values for D100 and BD100 fuels were observed to be very similar, attributable to their comparable cetane numbers. Under all test conditions, D100 and BD100 exhibited shorter ID durations, expressed in crankshaft angle degrees, compared to other fuels.

The ID period was extended when BD80E20, BD80P20, and BD80B20 fuels, prepared by mixing with biodiesels, were used. It was seen that the ID period was long, especially in the experiments conducted at 20% gas pedal position, and it was extended by 1.9 °CA compared to D100. The longest ID values were obtained as 14.8 °CA in BD80P20 and BD80B20 fuels.

In addition, when ID is evaluated in terms of SOI applications, it is seen that SOI increases ID in advanced applications. In addition, ID is shortened in retarded SOI applications in all fuel types. The extension of ID with the SOI in advanced timing applications is believed to result from the earlier fuel injection, which coincides with lower temperature and pressure conditions in the cylinder mixture. Since the fuel injection start is closer to TDC in the retarded SOI application, the temperature and pressure in the cylinder are higher than in the advanced SOI application. It is expected that ID will shorten as the fuel starts to burn in a shorter time.

The negative effect of low cetane numbers and high evaporation temperatures on the ID of alcohol-fueled blends was seen. The difference between the IDs decreases with increasing loads compared to single-fuel studies. Since the cetane numbers of biodiesel and diesel fuels are close to each other, similar IDs were observed under all conditions.

3.4. Combustion Duration

It is defined as the duration from the onset of fuel combustion in the cylinder until 90% of the injected fuel has been consumed, and it is also expressed in terms of crankshaft angle. Figure 9 shows the change in CD values at different pedal positions at different SOI timings. The CD was shortened for D100 fuel because of advancing and retarding SOI compared to the standard SOI application. In the experiments conducted at 20% gas pedal position, CD was shortened by 2.6 °CA in other SOI applications compared to the standard SOI application for D100 fuel.

The CD values of BD100 are close to D100 fuel but shorter. It is thought that the reason for the shortening of CD values in the use of BD100 fuel is due to the oxygen molecules in its content. In all test conditions, the longest CD is 58.5 °CA in D100 fuel at 80% pedal position and +5 °CA in advanced SOI application, while for BD100 fuel, the CD is 0.9 °CA shorter at 57.6 °CA under the same test conditions.

Additionally, when using fuels prepared with biodiesel–alcohol blends (BD80E20, BD80P20, and BD80B20), CD was observed to decrease by over 20% compared to D100 under identical test conditions. It is thought that the addition of alcohol fuel to biodiesel reduces the viscosity value due to its fuel properties, making the fuel more easily atomized and accelerating combustion. As ID lengthens, the quantity of fuel combusted during the uncontrolled combustion phase increases, while the fuel burned during the controlled combustion phase decreases. In the tests with alcohol mixtures, the CDs shorten with the prolongation of IDs. It is observed that maximum HRRs increase with the increase in IDs in alcohol fueled tests. In addition, the oxygen molecules contained in alcohol fuels accelerate the combustion process in the cylinder and cause CD to shorten. As the fuel quantity injected into the cylinder increased with higher GPP, the CD was extended.

3.5. CO Emission

CO emission is expressed as an exhaust emission caused by a lack of oxygen in the combustion process. Figure 10 shows the changes in CO values at different pedal positions in different SOI timings. The early SOI application significantly reduced CO emissions at 20% and 50% gas pedal positions. Compared to standard SOI applications for D100 fuel, advanced SOI applications decreased CO emissions by more than 10%. On the other hand, in 20% and 50% pedal applications, a deterioration of more than 5% was detected in CO emissions with retarded SOI when using D100 fuel. Early initiation of SOI is understood to facilitate a more uniform distribution of fuel within the combustion chamber by allowing increased mixing time, thereby enhancing combustion efficiency and resulting in a reduction in CO emissions.

In all test conditions, CO emissions were lower when using BD100 compared to D100. In particular, in tests conducted at 80% pedal position, a CO reduction of up to 25% was achieved when using BD100 compared to D100. When the use of biodiesel fuels containing alcohol was examined, CO emissions reached maximum release values under test conditions in the low pedal position and retarded SOI application. CO emission was measured at 20% GPP as 18.4 g/kWh for BD80E20, 25.2 g/kWh for BD80P20, and 29.9 g/kWh for BD80B20. When the CO emissions formed at low load are examined, it is seen that the CO emissions produced by alcohol mixtures increase. The positive effect of the oxygen contained in the alcohols could not eliminate the negative effect of low cylinder temperatures. However, by changing the injection timing to standard and then advanced, the in-cylinder temperatures increased, and CO emissions decreased. The rise in engine load causes an increase in in-cylinder temperatures, which in turn results in a reduction in CO emissions. Additionally, emissions recorded at 50% GPP demonstrated no statistically significant variation, indicating consistent combustion performance under this operational setting. However, a decreasing trend is observed due to the increasing engine load. At 80% pedal opening, the in-cylinder pressure and temperature rose due to the increased fuel quantity injected into the cylinder and the earlier timing of the injection. The effects of injection timings and engine load on cylinder pressures can be seen in Table 7. In the tests conducted at 80% pedal opening, low carbon and high oxygen contents ensured that alcohol blend fuels produced lower CO emissions.

3.6. CO₂ Emission

CO₂ emissions are released as a natural exhaust emission as a result of the combustion of fuels containing carbon. Figure 11 shows the changes in CO₂ values at different pedal positions at different SOI timings. At 50% GPP, there was an increase in CO₂ emissions with the advance of SOI for all fuel types. At 50% GPP, delayed CO₂ emissions for D100 fuel were measured as 728 g/kWh, while they were measured as 752.9 g/kWh for standard SOI and 807.3 g/kWh for advanced SOI. At 50% GPP, the highest CO₂ emissions for each fuel type with the advance of SOI were measured as 804.2 g/kWh for BD100, 762 g/kWh for BD80E20, 752 g/kWh for BD80P20, and 806.8 g/kWh for BD80B20. It is thought that the increase in CO₂ is due to improved combustion due to better distribution of the fuel in the cylinder and better mixing of the fuel–air mixture with the advance of SOI. Additionally, the highest CO₂ emissions were observed with 20% GPP under all test conditions. The highest CO₂ emissions were measured as 993.2 g/kWh with the delayed SOI application of D100 fuel and 20% GPP. The reason for the high CO₂ emissions at low GPP is thought to be due to the high BSFC at low GPP. Compared to D100, CO₂ decreased with the use of blended fuels. While CO₂ emissions were 792.5 g/kWh for D100 in the standard SOI application at 80% GPP, these were measured as 745.3 g/kWh for BD80E20, 743.6 g/kWh for BD80P20, and 791.9 g/kWh for BD80B20. The CO₂ reduction with blended fuels is due to their lower C/H compared to D100.

3.7. HC Emission

Hydrocarbon (HC) emissions represent the quantity of unburned fuel expelled from the cylinder during the exhaust phase. Figure 12 shows the changes in HC values at different pedal positions at different SOI timings. While the effect of SOI change on HC is clearly seen in the experiments conducted at 20% pedal position, it is seen that SOI change does not have much effect at 50% and 80% pedal positions. At 20% pedal position, advanced SOI caused a significant decrease in HC emissions compared to retarded SOI. While HC was measured as 0.1859 g/kWh in the retarded SOI application for D100 fuel, HC was obtained as 0.1453 g/kWh in the advanced SOI application, showing a decrease of more than 20%.

Furthermore, the use of BD100 fuel resulted in reduced HC emissions across all test conditions compared to D100 fuel. Improvements of more than 20%, 40% and 25% were observed at 20%, 50% and 80% pedal positions with BD100 fuel, respectively. The oxygen contained in biodiesel is thought to reduce HC emissions. However, the use of blended fuels at 20% GPP resulted in an increase in HC emissions of up to 25% compared to D100. This rise in HC emissions is attributed to the cooling effect of the alcohol components during evaporation, which adversely affects the combustion process within the cylinder.

On the other hand, the lowest HC emissions were measured as 0.04531 g/kWh for D100, 0.03395 g/kWh for BD100, 0.03354 g/kWh for BD80E20, 0.04139 g/kWh for BD80P20, and 0.04569 g/kWh for BD80B20 at 80% GPP. A decrease in HC emissions was observed with increasing GPP levels across all fuel types. Considering that the fuel accumulates and reaches the cylinder walls due to the high ignition delay at low GPP, it is in line with the expectations that HC is high at low GPP versus high GPP.

3.8. NO Emission

NO emissions are formed when nitrogen and oxygen atoms react under the high-temperature conditions generated during combustion within the cylinder. Figure 13 shows the changes in NO values at different pedal positions at different SOI timings. It was previously stated that high gas pressures are reached in the advanced SOI application. Accordingly, an increase in NO emissions was anticipated with advanced SOI. Upon analysis of the results, a significant rise in NO emissions was observed in advanced SOI conditions. Specifically, for D100 fuel, NO emissions more than doubled in advanced SOI applications compared to retarded SOI across all tested pedal positions. Conversely, under identical test conditions, BD100 fuel resulted in a reduction in NO emissions by over 20% relative to D100.

Maximum NO emissions in the experimental conditions were measured as 13.3 g/kWh, 12 g/kWh and 11.8 g/kWh with BD80P20, BD80E20, and BD80B20 fuels, respectively, at 50% pedal position and +5 advanced SOI application. This increase in NO emissions is believed to result from the elevated combustion temperatures caused by the high oxygen content present in alcohol-based fuels.

In the use of blended fuels, there was an increase in NO compared to D100 under all GPP conditions. It was determined that the increase in NO reached 26% with 20% GPP, 27% with 50% GPP, and 30% with 80% GPP with blended fuels. The blended fuels prepared by adding alcohol fuel to biodiesel have a better viscosity in terms of atomization, and due to the oxygen they contain, the fuel finds more oxygen in the cylinder, which results in an increase in NO.

4. Conclusions

In this study, the effects of five different fuel types (D100, BD100, BD80E20, BD80P20, and BD80B20) on combustion and exhaust characteristics were investigated in a single cylinder, four-stroke diesel engine with a common rail fuel injection system, at a constant engine speed of 2000 rpm and three different gas pedal positions (20%, 50% and 80%). As a result of the study, CGP, HRR, ID, and CD data were examined within the scope of combustion parameters, while CO, CO₂, HC, and NO data were examined within the scope of exhaust emissions.

It was determined that maximum CGP increases up to 25% with increasing GPP under standard operating conditions, and the maximum CGP values obtained were generally between 5 and 10 °CA after from TDC. It was observed that the maximum CGP was improved by more than 15% in all fuel types with advanced SOI values compared to retarded SOI values. Maximum HRR showed aggressive increases in biodiesel–alcohol fuel blends. The highest HRR value under standard operating conditions was obtained as 178.9 J/°CA in BD80P20 fuel at 80% GPP.

In advanced SOI, this generally resulted in an increase in ID for all fuel types. The highest ID values were observed at 20% GPP across all fuel types. The longest ID duration was 14.8 °CA for BD80P20 and BD80B20 fuels at 20% GPP with the advanced SOI. It was seen under the same experimental conditions that CD decreased in general when using the other four types of fuel compared to D100 fuel. The shortest CD was 32.2 °CA with BD80B20 in the retarded SOI application at 20% pedal position, while the longest CD was 58.5 °CA with D100 fuel in the advanced SOI application at 80% pedal position.

Experiments conducted at 20% and 80% pedal positions showed that advanced applications of SOI caused a reduction in CO emissions of more than 10% for all fuels compared to retarded SOI applications. The lowest CO emission under the experimental conditions was measured as 4.7 g/kWh in the BD80P20 with 80% pedal position and advanced SOI application, while the maximum CO was measured as 29.9 g/kWh in the BD80B20 with 20% pedal position and retarded SOI application.

Under standard conditions, the highest CO₂ emission values for all test fuels were measured at 20% GPP. As a result of using blend fuels prepared by adding alcohol to biodiesel, CO₂ emission values were found to be lower than D100 in all SOI applications.

The retarded SOI was found to adversely affect HC emissions. At 20% GPP, maximum CO emissions were measured for all fuel types with retarded application of SOI. The highest HC emission, measured at 0.3236 g/kWh, occurred with the BD80B20 fuel blend under retarded SOI at 20% GPP, whereas the lowest HC emission of 0.0306 g/kWh was recorded using the BD80E20 blend with advanced SOI at 80% pedal opening.

It was determined that NO emission was significantly affected by the SOI change. Advanced SOI applications increased NO emissions more than two times compared to standard SOI applications. It was observed that the use of biodiesel–alcohol blends led to an increase in NO emissions compared to D100, whereas the utilization of BD100 resulted in a reduction in NO emissions. In the experiments, the lowest NO emission, measured at 2.5 g/kWh, was recorded with BD100 under retarded SOI conditions at 50% pedal opening.

As can be seen from the data in this study, the results obtained with alternative fuels are close to those of D100, allowing its use without any modifications to the engine. In particular, it was observed that advancing the injection timing improved combustion parameters. SOI has been shown to have a significant impact on emission control, one of the factors hindering the use of diesel engines. However, future studies are also recommended to examine formaldehyde emissions resulting from the use of alcohol fuels. Also, it is recommended that future studies evaluate the cold flow properties of BD100 fuel, conducting longer tests under full-load engine conditions, and verifying these values through simulation.