Effect of Pilot Injection Strategy on Performance of Diesel Engine under Ethanol/F-T Diesel Dual-Fuel Combustion Mode

Abstract

To reduce emissions and save energy, alternative fuel and dual-fuel mode have been widely applied in the field of diesel engines. The pilot injection has potential to reduce engine vibration noise and pollutant emissions. The effects of a diesel fuel pilot injection strategy on the performance of an ethanol/F-T diesel dual-fuel engine were experimentally investigated on a four-cylinder four-stroke common rail diesel engine modified with an ethanol injection system. The results indicate that the variation in the combustion characteristic parameters with pilot injection timing is nonlinear and the difference is small, while soot, NO_x, and CO tend to decrease, with an increase in pilot injection timing. With the increase in pilot injection amount, p_max, combustion duration, CO and soot increased; p_max phase and CA50 were closer to TDC; HRR_max and the ignition delay period decreased. The BSFC tends to increase with the increase in pilot injection timing and the increase in pilot injection amount, while the BTE shows the opposite trend. The value and the variation range of COV_pmax are small. The effect of the pilot injection amount on ethanol/F-T diesel dual-fuel engine is more significant. The research presented in this paper can provide reference directions for the formulation of a fuel injection strategy of ethanol/F-T diesel dual-fuel combustion mode to reduce NO_x without worsening the combustion process and presenting an insufficient fuel economy.

1. Introduction

The diesel engine is widely used in agriculture, industry, transportation, and other fields. Due to the energy crisis, carbon emission requirements, and stringent emission regulations, as well as the complex environment surrounding oil in today’s world, energy security cannot be guaranteed; therefore, the research on diesel alternative fuels is very extensive [1]. NO_x and soot emissions caused by the typical combustion process of the diesel engine are a much-maligned issue that needs to be solved. Biomass-to-liquid [2], gas-to-liquid [3,4], and coal-to-liquid [5,6] alternatives, etc., are widely concerned in the study of diesel alternative fuels. China has few planned gas-to-liquid fuel projects due to a scarcity of resources [7]. In China, coal reserves are relatively large, and most coal-burning processes are not environmentally friendly, leading to a serious smog problem in the country. Using coal-to-liquid is one of the key measures to tackle the air pollution caused by the direct use of coal. According to China’s energy characteristics, fuel produced based on a clean utilization of coal is more favored [8].

F-T diesel is synthesized via the indirect liquefaction of coal using Fischer–Tropsch (F-T) technology [9]. Due to the addition of the coal gasification stage in the production process of F-T diesel, there is a significant difference between coal-to-oil and gas-to-liquid alternatives, including fuel composition, H/C ratio, distillation temperature, and viscosity, all of which play an important role in the formation of exhaust emissions [10]. The price equivalent of F-T diesel from coal is lower than the average crude oil price for the year of study in [11]. Reducing coal consumption, CO₂ emissions, and wastewater should be the most pressing issues explored in coal-to-liquid research [12]. Compared with petroleum-based diesel, F-T diesel has a higher H/C ratio, cetane number, and a lower sulfur content, and aromatic content; therefore, the combustion delay period of the engine can be shortened when F-T diesel is combusted in the cylinder, and the combustion temperature can be reduced at a certain extent to achieve low-temperature combustion. Moreover, F-T diesel can obtain lower NO_x, CO, and soot emissions [5,8,13]. In addition, F-T diesel fuel can be used in conventional diesel engines without any modification, with negligible or minimal engine efficiency gains [10]. F-T diesel can also be blended with other fuels to further improve engine performance. The study of Cai and Zhang et al. [14] showed that the blended fuel of F-T diesel and petrochemical diesel makes the engine run remarkably gently. NO_x emissions of the blends are only slightly reduced, and soot emissions are significantly reduced, resulting in the NO_x–soot trade-off being alleviated. Their study of blends of F-T diesel and gasoline shows that the addition of gasoline to F-T diesel can effectively improve brake thermal efficiency (BTE); however, CO, NO_x, and a concentration of ultrafine particulates at high loads are worsened [15]. Zhang et al. [16] found that when oxygenated fuels are mixed into the F-T diesel, the ignition timing is postponed, the ignition delay is prolonged, and the combustion duration is shortened. In conclusion, F-T diesel is a promising alternative fuel for compression–ignition (CI) engines that can reduce the dependence on petroleum fuels and reduce greenhouse gas emissions.

Alcohols are also a very promising alternative fuel with great sustainability. Alcohol fuels have been widely used as additives for petroleum-based fuels, which have great possibilities for energy independence, emission reduction, and energy resource sustainability [17,18,19]. In addition, the high latent heat of vaporization and lower flame temperature of alcohol fuels may be promising methods for increasing power output and reducing heat loss. Ethanol, due to its high oxygen content and absence of aromatic compounds, can reduce CO, NO_x, and soot emissions and has been extensively studied in CI engines. With low reactivity and cetane number, ethanol is usually used in CI engines via a blending process [20,21,22,23]. The results show that the performance and emission characteristics of diesel engines can be effectively improved using a blended fuel of diesel and ethanol. However, due to the presence of a hydroxyl group in anhydrous short-carbon chain alcohols, a molecular dipole moment is generated, resulting in poor insolubility between short-carbon alcohols and diesel oil, especially at low ambient temperature [24]. Therefore, when ethanol is used in the way of blending, the blended fuel requires an emulsifier, which has the disadvantages of high cost and poor low-temperature performance. In addition, the blended combustion mode cannot flexibly adjust the proportion of ethanol fuel with the operation to maximize the performance advantage of ethanol fuels; thus, the dual-fuel combustion mode of low-pressure injection of ethanol at the intake port becomes an interesting option [25,26].

The dual-fuel combustion mode, in which ethanol is injected at a low pressure in the intake port and diesel with high activity is injected directly in the cylinder, makes the ratio of ethanol more flexible than that of the blending mode [27]. Compared with the traditional diesel combustion mode, the dual-fuel mode can also significantly reduce soot and greatly reduce NO_x emissions, especially in the case of a high EGR [15,28]. Ethanol effectively prolongs the ignition delay period and simultaneously shortens the combustion duration because more premixed combustion takes place in the case of the dual-fuel combustion mode [29]. The dual-fuel mode can also reduce the maximum pressure in the cylinder and reduce CO and NO_x emissions [30,31]. In addition, Pedrozo et al. [31] showed that even if EGR was not used, the dual-fuel mode of ethanol and diesel could greatly reduce NO_x emissions and thus reduce the cost of the exhaust gas retreatment device. In short, the dual-fuel combustion mode can flexibly control the fuel injection ratio according to the requirements of the operating conditions so that the engine has better performance.

Engine performance may be further improved through the control of the injection of in-cylinder direct injection diesel [32]. The dual-fuel combustion of pre-mixed ethanol fuel ignited by diesel from a single-stage injection limits the operating range of the engine due to the high pressure rise rate (PRR). The multi-stage injection of the diesel engine can significantly reduce the PRR in the cylinder, thus reducing the vibration and noise in the working process. In addition, NO_x and soot emissions can be improved by optimizing the multi-stage injection control parameters [33,34]. The diesel injection mode with the pilot injection stage can expand the working range and reduce PRR. In addition, during advanced combustion, NO_x emissions increase and soot emissions reduce due to the higher pressure and temperature in the cylinder, which leads to soot oxidation [31]. Pan et al. [35] showed that the cyclic variations of ethanol/diesel dual-fuel combustion are sensitive to the diesel injection timing. With the increase in diesel injection timing, the coefficient of the variations (COVs) of all parameters decreases at first and then increases; and the distribution ranges of the crank angle of the maximum in-cylinder pressure (p_max), the maximum heat release rate (HRR_max), and the maximum PRR (PRR_max) are narrowed firstly and then widened. Luo et al. [36] pointed out through a simulation study that the proper main and post-injection strategy can improve the combustion and emission characteristics, especially the reduction in NO_x and CO emissions.

In conclusion, while using clean alternative fuels, it is possible to achieve better engine performance and cleaner emissions through the optimization of the engine parameters in order to achieve lower cost energy and a reduction in emissions. Although the main injection parameters have a greater impact on combustion and emissions in general, the pilot injection stage with a relatively small amount of fuel also has a great impact on fluid-dynamic conditions at the start of the main injection and fuel evaporation and combustion in the main injection stage, resulting in a reduction in PRR and NO_x. A reasonable pilot injection timing can simultaneously reduce NO_x and soot emissions, and fuel consumption, as well as improve the economy of the engine, which is the advantage of multi-stage injection. In this paper, the influence of pilot injection parameters on the engine performance of the dual-fuel combustion mode of F-T diesel and ethanol under the condition of fixed ethanol substitution rate (ESR) and operation conditions is studied experimentally.

2. Experimental Setup and Methodology

2.1. Testing Bench and Equipment

The engine experiments were conducted on a four-cylinder, four-stroke, direct-injection diesel engine to examine the combustion and exhaust emissions. The engine model is D25 and is made by Kunming Yunnei Power Co., Ltd. (Kunming, China). The engine is equipped with a high-pressure common-rail injection system that enables a multi-stage F-T diesel injection process. Compared to a single injection strategy, the use of pilot injection can extend the engine’s operating range and reduce PRR while having minimal impact on the soot/NO_x trade-off [31]. The pilot injection timing (t_pil) and pilot injection amount (q_pil) of the original engine in the experimental condition are 12.3 °CA before the top dead center (BTDC) and 1.5 mg per cycle (mg/cyc), respectively. The specifications of the diesel engine are presented in

Table 1. Specifications of the test engine before modification.

Items	Values
Cylinder number	4
Bore (mm)	92
Stroke (mm)	94
Displacement (L)	2.499
Compression ratio	16.6:1
Rated power speed (r·min−1)	3200
Rated power (kW)	85

. To realize a dual-fuel combustion mode, four ethanol injectors are added to the intake manifold of the diesel engine for the multi-point injection of ethanol. The injection parameters were controlled using an open-source integrated dual-fuel electronic control unit (ECU).

The schematic of the experimental setup is shown in Figure 1. The engine speed and torque are controlled through the ET4000 engine control system. A precision crank angle (CA) encoder was connected to the engine crankshaft to determine the accurate position of the piston concerning crank angle. A Kistler 6125B cylinder pressure sensor and a Kistler combustion analyzer (Kistler, Winterthur, Switzerland) were used for in-cylinder pressure acquisition and heat release analysis. The number of in-cylinder pressure signals collected in each operating cycle is 7200. The measured in-cylinder combustion pressures at different crank angle positions are used to calculate the heat release rate. Gaseous emissions are measured with the multi-component exhaust emission analyzer (AVL SES-AM i60 FT, AVL, Jinan, China). Soot emissions are measured with the micro soot emission test system (AVL Micro Soot Sensor 483). For the operating range and accuracy of the test equipment, see the previously published article reference [37]. The accuracy values of the test equipment are presented in

Table 2. Operating ranges and accuracies of the test equipment.

Measured Parameters	Unit	Range	Accuracy
Engine speed	r·min−1	0~8000	±0.1% F.S
Engine torque	N·m	0~500	±1% F.S
Cylinder pressure	bar	0~250	±0.005
F-T diesel mass flowmeter	kg/h	0~250	0.2% F.S
Ethanol mass flowmeter	kg/h	0~40	0.1%
NOx	10–6	0~10,000	±1.5
CO	10–6	0~10,000	±1.2
Soot	mg/m3	0.001 to 50	±0.005

.

The combustion characteristics of diesel engines are affected by two important parameters, that is, the cetane number and the oxygen content of the test fuel. Two fuels were used: ethanol (the low-reactivity fuel) and F-T diesel (the high-reactivity fuel). The physical and chemical properties of the two test fuels are shown in

Table 3. The properties of the fuels.

Parameters	0# Diesel [13]	F-T Diesel [37]	Ethanol
Density at 25 °C (g/cm3)	0.81	0.757	0.789
Cetane number	55.8	74.8	6
Research octane number	-	-	111
Low heating value (MJ/kg)	42.65	43.07	26.79
Heat of vaporization (kJ/kg)	260	-	847
Oxygen fraction (mass%)	0	0	34.8
Stochiometric air/fuel ratio	14.3	14.96	9.0

. Compared to the density and low heating value (LHV) of petroleum-based diesel, the volume energy density of F-T diesel is lower, which leads to more fuel needing to be injected into the cylinder to achieve the same power as that under the original engine parameters.

2.2. Methodology

The experimental study was performed under the speed of 2000 r·min⁻¹, which is the speed that corresponds to the maximum torque and the load factor of 75% (168 N·m). The engine’s fuel temperature is set to 20 °C, while the coolant’s temperature and lubricating oil’s temperature are set to 90 °C.

HRR is calculated via the pressure in the cylinder based on the application of the first law of thermodynamics in the closed system and the equation of state under the condition that the working medium in the cylinder is regarded as an ideal gas. As a result of it omitting heat transfer to walls, crevice volume, blow-by and the fuel injection effect, among others, HRR is termed the net HRR, and calculated as in Formula (1) [38].

$$\mathrm{HRR}=\frac{1}{\gamma -1}\left(\gamma p\frac{\mathrm{d}V}{\mathrm{d}\phi }+V\frac{\mathrm{d}p}{\mathrm{d}\phi }\right)$$

(1)

where γ is the specific heat ratio, p is the instantaneous pressure in the cylinder measured using the cylinder pressure sensor in Pa, and V is the instantaneous cylinder volume in m³.

A simple diagram of the dual-fuel mode injection strategy used in this test is shown in Figure 2. Ethanol is injected into the intake port at low pressure, while the intake valve is open, and enters the cylinder through the intake valve. The F-T diesel enters the cylinder via direct injection before the top dead center (TDC), which includes a multi-stage injection process of the pilot injection stage and the main injection stage. The fuel distribution was based on the circulating fuel mass and the main injection fuel mass being the difference between the total fuel mass and the pilot injection fuel mass. The injection control parameters of F-T diesel are shown in

Table 4. The pilot injection parameters of F-T diesel.

Parameters	qpil (mg/cyc)	tpil (°CA BTDC)	Main Injection Timing (°CA BTDC)	Period between the End of Pilot Injection and the Start of Main Injection (°CA)
Fixed qpil	1.5	10.3	5.6	1.8
		12.3		3.8
		14.3		5.8
		16.3		7.8
		18.3		9.8
Fixed tpil	1.5	12.3	5.6	3.8
	2.0			3.7
	2.5			3.6
	3.0			3.5
	3.5			3.4
	4.0			3.3

.

The ESR is used to calculate the amount of ethanol addition in the intake port to realize the dual-fuel combustion mode. The calculation formula of the ESR is shown in Formula (2):

$$\mathrm{ESR}=\frac{{\dot{m}}_{\mathrm{E}}\cdot h_{\mathrm{E}}}{{\dot{m}}_{\mathrm{E}}\cdot h_{\mathrm{E}}+{\dot{m}}_{\mathrm{FT}}\cdot h_{\mathrm{FT}}}\times 100\%$$

(2)

where ${\dot{m}}_{\mathrm{E}}$ is the ethanol injection rate in kg/h, ${\dot{m}}_{\mathrm{FT}}$ is the F-T diesel injection rate in kg/h, h_E is the LHV of ethanol, and h_FT is the LHV of F-T diesel in MJ/kg. Compared with the traditional combustion mode, the dual-fuel combustion mode combined with EGR can achieve low-temperature combustion and significantly reduce NO_x emissions, as shown in Figure 3. At the same time, although there was an increase in soot, CO and sum non-methane hydrocarbon compounds (NMHC) did not increase when EGR was used. Therefore, this paper studies the effect of the pilot injection strategy of highly active fuel with direct injection in the cylinder on the combustion and emission characteristics of the ethanol F-T diesel dual-fuel mode under the condition that the EGR is 15% and the ESR is 20%.

The equivalent brake specific fuel consumption (BSFC) represents the total BSFC calculated after converting LHV such as ethanol consumption rate into highly active fuel under the dual-fuel mode; the unit is g/(kW·h), and the calculation formula is shown in Formula (3):

$${\mathrm{BSFC}}_{\mathrm{equivalent}}=\frac{m_{\mathrm{FT}}+\frac{h_{\mathrm{E}}}{h_{\mathrm{FT}}}{m}_{\mathrm{E}}}{P_{\mathrm{e}}}\times {10}^3$$

(3)

where P_e is the engine brake power, kW.

Since the combined combustion of two fuels is involved in the test, it is extremely inaccurate to measure the BTE of a single fuel. Therefore, the BTE of the dual-fuel engine is calculated by converting the LHV of the two fuels into F-T diesel consumption, and the formula is shown in Formula (4):

$$\mathrm{BTE}=\frac{3600}{{\mathrm{BSFC}}_{\mathrm{eequivalent}}\cdot h_{\mathrm{FT}}}$$

(4)

Engine in-cylinder pressure is the most important parameter for analyzing factors affecting engine performance because it contains useful information about what happens in the engine chamber. The combustion HRR can be calculated from the cylinder pressure data based on the first law of thermodynamics. Instantaneous cumulative heat release (CHR) percentage is often used to segment combustion events.

The crank angles corresponding to the CHR reaching 5% (CA05) and 90% (CA90) are often considered the combustion’s starting point and the combustion’s ending point, respectively. Thus, the combustion duration can be calculated. In addition, the crank angle corresponding to the CHR reaching 50% (CA50), known as the center of combustion, is also very important to the analysis of the combustion process. The ignition delay can be calculated by the crank angle from the main injection timing to CA05.

Many pressure-related parameters can be used to study cyclic variations. In this paper, the coefficient of variation (COV) of maximum in-cylinder pressure (p_max) is used to assess the cyclic variations of the ethanol/F-T diesel dual-fuel combustion mode. It can be calculated as in Formula (5):

$${\mathrm{COV}}_{p_{\max }}=\frac{\sqrt{{\displaystyle \sum _{i=1}^N{\left(p_i-\overline{p}\right)}^2\cdot \frac{1}{N}}}}{\overline{p}}$$

(5)

where p_i denotes the p_max per cycle, N represents the engine operating cycle number (N is 200 in the study), and $\overline{p}$ is the mean value of the p_max for N cycles.

3. Results and Discussion

3.1. Combustion Characteristics

The in-cylinder pressure data in the crank angle of 200 consecutive engine cycles were collected, and the average dataset was then used to calculate combustion characteristics parameters, such as HRR, combustion duration, ignition delay period and cycle fluctuation, etc.

Figure 4 shows the in-cylinder pressure curves and HRR curves with crank angles under different pilot injection timings (Figure 4a) and pilot injection amounts (Figure 4b). The HRR curves show a trend of multi-stage heat release. The pilot injection strategy has a significant influence during the early combustion process, and then indirectly affects the heat release law in the middle and late combustion process.

As can be seen from Figure 4a, the HRR phase in the pilot injection stage changes obviously, while the p_max and HRR_max in the main injection stage show little difference. However, when the pilot injection timing is 10.3 °CA BTDC, it is too close to the TDC, which makes the interval between the pilot injection and the main injection too short, and the boundary between the combustion stage is not obvious. The PRR_max of the main combustion stage is too high, resulting in rough engine operation and the inability to achieve low-temperature combustion. It can be seen from Figure 4b that the HRR_max value of the pilot injection stage is located near 3 °CA BTDC. With the increase in the pilot injection amount, the HRR_max of the pilot injection stage increases and the heat release amount of the main injection stage reduces. In addition, the p_max increases gradually with the increase in the pilot injection amount. This is because with the increase in pilot injection amount, the amount of energy used to ignite the in-cylinder ethanol homogeneous mixture increases, causing more pre-mixed F-T diesel fuel to be ignited.

To analyze the combustion process parameters in further detail, the parameters of p_max, p_max phase, HRR_max, COV_pmax, CA50, ignition delay period, and combustion duration were extracted, as shown in Figure 5.

Figure 5a shows the influence of the pilot injection timing on the combustion parameters. The increase in pilot injection timing leads to a sufficient amount of time for the pilot injection F-T diesel to atomize and vaporize in the cylinder, making the fuel and air mix more full. A thermos atmosphere is formed in the cylinder before the main injection stage. However, with the increase in pilot injection timing, the influence of pilot injection on the main injection stage may be weakened. It can be seen from Figure 5(a1–a3) that HRR_max and p_max have the same variation trend, along with the pilot injection timing, and the minimum values are obtained near 14.3 °CA BTDC, while the p_max phase delays slightly. When the pilot injection timing is closer to the TDC, the thermos atmosphere generated by the pilot injection in the cylinder has a stronger effect on the heating and evaporation of the fuel in the main injection stage, and the ignition delay period in the main injection stage is slightly shorter, which makes the p_max and the HRR_max slightly increase, and CA50 is slightly close to TDC (Figure 5(a5–a7)).

Figure 5b shows that the influence law of pilot injection amount on the combustion process parameters at the pilot injection timing of 12.3 °CA BTDC is changed. As can be seen from Figure 5(b1–b3), the peak value of HRR_max at the pilot injection stage increases with the increase in pilot injection amount; the p_max also increases gradually, while HRR_max of the main injection stage decreases continuously. The phase of p_max is closer to TDC with the increase in the pilot injection amount.

On the one hand, with the increase in the pilot injection amount of F-T diesel, the combustion chamber will accumulate a large number of active free radicals and oxidation reactive substances, which weaken the inhibition effect of ethanol on the ignition of F-T diesel, increase the possibility of simultaneous ignition of multiple points, and enhance the oxidation and reaction rate of the combustion process, resulting in the shortening of the ignition delay period of the main injection stage (Figure 5(b6)). Due to the shortening of the ignition delay period in the main injection stage, the F-T diesel in the main injection stage will participate more in diffusion combustion under the current injection parameters, and the proportion of exhaust gas in the cylinder during the main injection stage is increased due to the large amount of fuel burned in the pilot injection stage; therefore, the instantaneous HRR decreases significantly, HRR_max decreases, and the combustion duration is prolonged, as shown in Figure 5(b7).

On the other hand, when the fuel amount in the main injection stage is reduced, the diffusion combustion ratio of the whole combustion process is reduced, the average burning rate increases, the combustion is more concentrated, and the CA50 is closer to the TDC (Figure 5(b5)); therefore, the p_max increases. However, the increase in the combustion phase may result in higher wall heat transfer losses, which will be detrimental to the engine’s efficiency.

As can be seen from Figure 5(a4,b4), the COV_pmax fluctuates slightly, decreasing with the increase in pilot injection timing, and increasing with the increase in pilot injection amount. However, the stability of the overall engine working cycle is within the acceptable range. It shows that the multi-stage injection strategy reduces the inhomogeneous fuel/air mixture in the cylinder and enhances the combustion stability of the engine. The addition of ethanol leads the premixed combustion ratio to be higher; therefore, the combustion stability of the dual-fuel engine is better.

3.2. Fuel Economy Performance

Although the research in this paper is aimed at constant engine operating conditions, the adjustment of pilot injection parameters will inevitably affect the combustion efficiency, and then affect the heat release and work capacity of the same fuel amount. Therefore, it is necessary to evaluate the influence of pilot injection parameters on the engine combined with fuel consumption.

Figure 6 shows the variation rules of BSFC_equivalent, BSFCs of F-T diesel and ethanol under different pilot injection parameters. As can be seen from Figure 6a, the BSFC of F-T diesel is basically constant. With the increase in the pilot injection timing, ethanol consumption increases, and the equivalent BSFC increases. The pilot injection time of 12.3 °CA BTDC shows a better fuel economy.

As can be seen from Figure 6b, the BSFC of F-T diesel under the current working condition is also essentially constant, while the ethanol consumption increases with the increase in the pre-injection amount, and the equivalent fuel consumption rate registers little change. It can be concluded from Figure 6 that the fuel economy is better when the pilot injection timing is 12.3 °CA BTDC and pilot injection amount is 1.5 mg.

Ethanol is injected into the combustion chamber via a low-pressure injection in the intake manifold port. During valve overlap, some of the premixed ethanol is expelled directly from the engine, which reduces the BTE. On the other hand, premixed ethanol will absorb and recover part of the engine block heat during the intake and compression stroke, and reduce the intake temperature to increase the intake charge, which is conducive to the BTE of the engine (especially under heavy load conditions). In addition, the addition of ethanol will prolong the ignition delay period, delaying the combustion phase (as shown in Figure 5(a5,a6)), and at the same time make the combustion more inclined to premixed combustion. The competition between the two determines the HRR phase and affects the BTE of the engine. The changing trend of BTE with the pilot injection parameter is shown in Figure 7. The BTE and BSFC of the engine are inversely proportional when the LHC of the fuel is constant.

As can be seen from Figure 7a, the BTE of the engine firstly increases and then decreases with the increase in the pilot injection timing, and reaches a maximum value at 12.3 °CA BTDC. Compared with the original parameters, with the increase in the pilot injection timing, the temperature in the cylinder at t_pil decreases, and the BSFC of ethanol significantly increases although the ESR is constant, which leads to the BTE decreasing obviously. It can be seen from Figure 7b that compared with the original engine parameters, the BTE decreased significantly with the increase in the pilot injection amount. Similarly, as can be seen from Figure 6b, the BSFC of ethanol increases significantly with the increase in the pilot injection amount. In this case, the lower temperature in the cylinder may lead to a more incomplete combustion of F-T diesel fuel, resulting in a decrease in the final thermal efficiency.

3.3. Exhaust Emission Analysis

Due to the high compression ratio, the high combustion temperature in the cylinder, and the large equivalent air-fuel ratio of the diesel engine, the nitrogen in the cylinder will be more oxidized to NO_x than that in the gasoline engine. Soot is produced due to partial oxidation, the thermal cracking of fuel under high temperatures, and a rich fuel mixture. CO is the main intermediate product produced in the combustion process of hydrocarbon fuels. If the oxygen concentration and temperature of the mixture in the cylinder are high enough, and the time occupied by the chemical reaction is long enough, CO will be oxidized into CO₂. Figure 8 shows the influences of the pilot injection parameters on soot, NO_x, and CO emissions in ethanol/F-T diesel dual-fuel mode.

As shown in Figure 8a, the variation law of emissions with the pilot injection timing is nonlinear; however, pollutant emissions decreases with the increase in the pre-injection timing overall. Especially at the pilot injection timing of 18.3 °CA BTDC, all three emissions are smaller because the fuel during the pilot injection stage has enough time to form a homogeneity mixture, which burns more fully. Soot and NO_x emissions show a significant trade-off relationship under a pilot injection timing of 14.3 °CA BTDC, which indicates that the combustion temperature in the cylinder is simultaneously lower. It can also be seen from Figure 5(a1) that the p_max at this pilot injection timing is the minimum. Compared with the pilot injection timing of 14.3 °CA BTDC, the pilot injection timing of 18.3 °CA BTDC will lead to a higher temperature and pressure in the cylinder at the end of the compression stroke.

As shown in Figure 8b, the variation law of emissions with pilot injection amount is also nonlinear; however, soot and CO emissions increase with the increase in pilot injection amount overall. NO_x emissions decrease significantly at the pilot injection amount of 2.5 mg but have little change at other pilot injection amount parameters. With the increase in the pilot injection amount, the ignition delay period of the main injection stage is shortened, as shown in Figure 5(b6), which results in a greater proportion of the main injection stage fuel involved in diffusion combustion. The mixture of fuel and air is relatively inhomogeneous, and the local rich fuel region increases; thus, soot and CO emissions tend to increase.

In conclusion, under the condition of slightly increased pilot injection timing and slightly increased pilot injection amount, significantly lower NO_x emissions can be achieved, which will provide research directions for further improving the performance of ethanol/F-T diesel dual-fuel engines.

4. Conclusions

The effects of the pilot injection strategy on the combustion and emission characteristics of ethanol/F-T diesel dual-fuel diesel engines were investigated. The potential of the ethanol/F-T diesel dual-fuel mode was explored. The main conclusions are as follows:

(1)

With the increase in pilot injection timing, the change in the combustion process parameters is nonlinear and the difference in the parameters is small. When the pilot injection timing is 14.3 °CA BTDC, the maximum p_max and the minimum HRR_max are obtained. With the increase in pilot injection timing, COV_pmax tends to decrease and the BSFC_equivalent tends to increase, while BTE tends to reduce. The best engine economy is obtained when the pilot injection timing is 12.3 °CA BTDC. With the increase in pilot injection timing, soot, NO_x, and CO showed a decreasing trend.

(2)

With the increase in the pilot injection amount, p_max, combustion duration, CO, and soot increased; the p_max phase and CA50 are closer to TDC; and the HRR_max and ignition delay period are reduced. COV_pmax and the BSFC_equivalent tend to increase, while BTE tends to reduce. When the pilot injection amount is 2.5 mg, NO_x emissions decreased significantly.

In summary, in the case of constant ESR and operation conditions, the adjustment of the pilot injection parameters can achieve a significant reduction in NO_x emissions without worsening the combustion in the cylinder and without a great reduction in fuel economy. This study could serve as a guide for the future application of coal-based alternative fuels in dual-fuel mode, which could be friendly for the environment.