Effects of Pre-Injection Strategy on Combustion Characteristics of Ammonia/Diesel Dual-Fuel Compression Ignition Mode

Abstract

As a zero-carbon clean fuel, the use of ammonia in internal combustion engines is of great significance to achieve the “two-carbon” goal. This paper investigates the effect of the ammonia energy ratio and diesel injection timing on combustion and emissions. Based on Computational fluid dynamics (CFD) and the simulation analysis software, Converge, three-dimensional modeling is carried out for a direct injection diesel engine with a Compression Ignition (CI) mode. Under the initial full-load conditions of 1200 r/min, the engine simulation was calculated. The results show that the peak cylinder pressure increases and then decreases as the ammonia energy ratio increases, the ignition delay time increases, and the CO₂ and N₂O emissions decrease. With pre-injection, the peak cylinder pressure increases at the same energy ratio and the combustion stage advances, resulting in improved indicated thermal efficiency. In comparison to the pure diesel mode, the pre-injection strategy shows an obvious reduction in greenhouse gas (GHG) emissions with a decrease of 40.9% by adjusting the injection timing, while the single injection strategy shows a reduction of 36.5%. The soot emission peak occurs in the diesel-only mode with 98.13% and 99.6% reductions in emissions under single and pre-injection, respectively. The ammonia–diesel dual-fuel (ADDF) engine with an ammonia-to-energy ratio of 70% and optimized ammonia and diesel injection timing significantly reduces the NH₃ emissions and GHG emissions by 69.34%.

1. Introduction

With global warming and increasingly stringent emission regulations [1], the use of clean alternative fuels for internal combustion engines has become a top priority for the internal combustion engine industry. Ammonia (NH₃) is a zero-carbon fuel that produces only clean, pollution-free water and nitrogen when completely burned. Ammonia is a colorless, flammable gas at room temperature, with a pungent odor, a high octane number, good knock resistance [2], high energy density, easy storage and transportation [3], perfect industrial infrastructure, and a low production cost [4]. Due to the reduction in greenhouse gas (GHG) emissions, the application of ammonia in internal combustion engines has received a lot of attention [5].

In order to achieve the dual-carbon goal, the application research of an ammonia–diesel dual-fuel (ADDF) engine is increasing year by year. The ADDF engine modification enables engine decarbonization at a low cost. It is a small modification to the original equipment. And green ammonia as an important technical choice for the internal combustion engine to achieve zero carbonization; its practical application needs to solve the ammonia internal combustion engine ignition difficulty, slow down combustion, NO_x and NH₃ emissions, and power reduction, and other technical problems. In addition, ammonia produces N₂O during the combustion process, and N₂O has an extremely strong greenhouse effect, which is nearly 300 times that of CO₂, and it can offset the reduction in CO₂ emissions, so N₂O emission is a special concern for ADDF engines, and GHG emissions are evaluated based on CO₂ and N₂O in this study.

Research scholars are optimistic about the prospects of ammonia as a zero-carbon fuel, and there is a growing body of research on ADDF applications. Studying the low-pressure dual-fuel (LPDF) and high-pressure dual-fuel (HPDF) injection modes of the ADDF engine, Li et al. [6] discovered that the HPDF mode has an advantage over the LPDF mode in terms of reducing unburned NH₃, NO_x, and GHG emissions. Yousefi et al. [7] investigated the effects of the ammonia energy fraction and diesel injection timing on the combustion and emission performance of an ADDF engine and showed that increasing the ammonia energy fraction decreases the thermal efficiency, increases N₂O emissions, and results in higher GHG emissions compared to diesel-only combustion. Yousefi et al. [8] also investigated the effects of diesel segmental injection on the combustion and emissions of the ADDF engine. The study revealed that the addition of ammonia to single-injection diesel decreased the indicated thermal efficiency (ITE). However, in the case of multiple-injection diesel, the ITE of the ADDF combustion mode was higher than that of the diesel-only mode, which reduced unburned NH₃ and GHG emissions. Liu et al. [9] investigated the performance of engines using liquid and gaseous ammonia fuels, and the results showed that ammonia engines, whether fueled with liquid or gaseous ammonia, can achieve the load range and thermal efficiency of a marine engine, and it was found that liquid ammonia outperformed gaseous ammonia in terms of power, ammonia emissions, and thermal efficiency. Zhang et al. [10] investigated the effect of diesel injection timing on the combustion of a two-stroke, low-speed liquid ammonia–diesel dual direct injection engine, and the results showed that the ADDF mode improves the ITE, the engine ignition phase is controlled by the diesel injection timing, and the reduction in fuel-based NO_x emission can be achieved by adjusting the injection timing of diesel and ammonia. Jin et al. [11] investigated the effect of the injection strategy on the combustion of a gaseous ammonia engine with low-pressure injection and found that the overall ITE can be improved and the GHG and ammonia emissions can be reduced by optimizing the injection strategy.

At present, ADDF engine research mainly focuses on gaseous ammonia injection in low-intake and marine two-stroke engines, while the combustion strategy of the direct injection of liquid ammonia in ADDF for diesel ignition has not yet been perfected. High-pressure liquid ammonia direct injection can not only provide enough ammonia to the cylinder in a short time, but also offers a more flexible injection strategy. In addition, it is an efficient method to guarantee engine power and enhance the energy replacement rate of ammonia fuel. Therefore, it is worthwhile to develop a high-pressure liquid ammonia for in-cylinder ADDF.

In this study, we investigated the ADDF combustion mode of a modified four-stroke off-road diesel engine using CFD. We analyzed the engine’s combustion and emission characteristics under various ammonia energy ratios and diesel injection strategies.

2. Research Models and Methods

2.1. Engine Parameters

In this study, a certain in-cylinder dual-fuel direct injection engine is used as a research model. The original machine is equipped with a high-pressure common rail fuel injection system located on top of the cylinder and features three cylinders. The other main parameters are presented in

Table 1. Main parameters of original machine.

Item	Parameter
Combustion chamber form	ω
Bore (mm)	78
Stroke (mm)	85
Rated power (kW/(r/min))	18.4/2600
Max. torque (N∙m/(r/min))	74.2/1820
Compression ratio	19:1
Displacement/L	1.218

. The model utilizes in-cylinder direct injection for all injection methods, with liquid ammonia fuel injected prior to the diesel fuel. Upon reaching the ignition limit at the end of the compression stroke, the diesel fuel undergoes compression ignition, leading to the simultaneous combustion of the surrounding NH₃-air mixture, forming a mixed combustion state.

2.2. Simulation Model Setup

To investigate how pre-injection strategies impact the ADDF combustion process, we employed a method that couples chemical reaction kinetics with a computational fluid dynamics (CFD) model. After establishing an operational model, we simulated it using the fluid simulation software, Converge. The reduction chemistry of ammonia/n-heptane mixtures was examined using the chemical kinetic reaction mechanism proposed by Wang et al. [12]. This mechanism comprises 495 reactions and 74 species. The combustion calculation models [13,14,15,16,17,18,19,20,21] were selected as shown in

Table 2. Selection of combustion calculation model.

Physical Process	Model
Turbulence model	RNG k-ε [13]
Spray breakup	KH + RT [14]
Droplet collision	NTC [15]
Spray/wall interaction model	Wall film [16]
Turbulent dispersion	O’Rourke [17]
Evaporation	Frossling [18]
Combustion model	SAGE [19]
NOx emission model	Extended Zeldovich NOx [20]
Carbon smoke emission model	Hiroyasu soot [21]

. The calculation model included the intake, exhaust, and cylinder, as shown in Figure 1. The simulation calculations focus on the combustion process of an ADDF engine with the rotation angle of the crankshaft between opening the intake valve and closing the exhaust valve. The Converge v2.4 software’s automatic mesh generation technology, combined with multiple mesh control techniques, was used to set the basic mesh size to 4.0 mm for the model. The simulation results were then obtained through adaptive encryption. After verification, this size satisfied the calculation requirements.

2.3. Simulation Model Verification

The engine bench test was conducted on the prototype machine of the simulation model, and the bench test system included a diesel high-pressure common rail fuel injection system, an ammonia injection system, intake and exhaust systems, and a combustion analysis system. The schematic diagram of the bench test platform is shown in Figure 2a. In the test, the Sichuan Polis ET2000 measurement and control system was adopted. The combustion pressure in the combustion chamber was measured using Kistler’s 6052C cylinder pressure sensor, and the cylinder pressure data were collected and analyzed using Devetron’s DEWE-800-CA-SE combustion analyzer. The crankshaft angle resolution was set to 0.1° CA, the engine operating condition was set to n = 1200 r/min, and IMEP = 0.5 MPa. Figure 2 shows the comparison between the in-cylinder pressure and the heat release rate measured in the bench test and the simulation calculations. In Figure 2b, it can be seen that the in-cylinder pressure test values measured using the original machine generally match the changes in the simulated values obtained from the model simulation. The cylinder pressure peak occurs at a crankshaft angle (CA) difference of 0.3° with a cylinder pressure peak error of 0.7%, indicating that the simulation model that was built is reasonable.

2.4. Simulation Research Scheme

In this study, the simulation conditions comprised a rotational speed of 1200 r/min and a 100% load. The main research contents include the following parts: Firstly, this study investigated the effect of the ammonia energy ratio on the ADDF engine operating under single-injection and pre-injection strategies for diesel. The ammonia energy ratio was calculated as the ratio of the heat value of ammonia to the total heat value of fuel in the cylinder. The physico-chemical properties of the fuel are shown in

Table 3. Fuel properties [7,22,23].

Fuel Properties	Ammonia	Diesel
Research octane number	130	-
Boiling point (°C)	−33.4	180–360
Ignition point (°C)	800	220
Minimum ignition energy (MJ)	680	0.63
Low heating value (MJ/kg)	18.8	42.5
Laminar flame speed (cm/s)	10	33
Auto-ignition temperature (K)	930	527–558
Adiabatic flame temperature (K)	2073	2573

. When studying pre-injection, it was 50% of the total diesel injection. Secondly, this study investigated the effect of diesel injection timing on the combustion process of the ADDF engine utilizing the pre-injection strategy. The injection timing for engine combustion was tested based on the parameters specified in serial cases 3 and 4.

Table 4. Numerical simulation research conditions of dual-fuel combustion mode.

Case	Speed/r/min	Ammonia Injection Timing/° CA ATDC *	Diesel Pre-Injection Timing/° CA ATDC	Diesel Main Injection Timing/° CA ATDC	Ammonia Energy Ratio/%	Injection Pressure/MPa
1	1200	−60	-	−6	0/20/40/60/65/70/75/80	58.8 (diesel), 60 (ammonia)
2			−26
3		−150	−26	−14/−12/−10/−8/−6/−4	70
4			−50/−42/−34/−26/−18	−6

* ATDC (After Top Dead Center).

shows the actual numerical simulation research parameters employed.

3. Results and Discussions

3.1. Effects of Ammonia Energy Ratio on Combustion and Emission

3.1.1. Effects of Ammonia Energy Ratios on Combustion Characteristics

Figure 3 illustrates the impact of the ammonia energy ratio on the in-cylinder pressure when utilizing the single-injection and pre-injection strategies for pilot diesel fuel at full-load and 1200 r/min conditions. From the figure, it is evident that the peak value of the in-cylinder pressure displays a rising and then decreasing tendency under these two injection strategies as the ammonia duty cycle increases, with its corresponding crankshaft angle showing a continuous backward trend. Due to the fuel characteristics of ammonia, ammonia cannot ignite spontaneously in the engine and can only be ignited by diesel oil. When the energy ratio of the ammonia increases, the diesel content within the combustion chamber decreases, which reduces the combustion speed of the engine and ultimately results in the decrease in the maximum combustion pressure. Increasing the energy ratio of ammonia will increase the ignition delay and premixed combustion stage and reduce the diffusion combustion stage, so the cylinder pressure under an energy ratio of less than 70% will exceed the diesel-only mode, while the latent heat of ammonia vaporization and the ignition delay effect will be enhanced when there is an energy ratio of more than 70%, so the cylinder pressure will be lower than that in the diesel-only mode. It can be observed from Figure 3 that the engine misfires when the ammonia ratio reaches 80%. As a result, the upper limit of the ammonia ratio for this engine is 75%. At the same time, when comparing the single-injection strategy shown in Figure 3a to the pre-injection strategy shown in Figure 3b, the peak cylinder pressure at the same ammonia ratio is increased with the latter. The pilot diesel pre-injection promotes the full atomization and evaporation of pilot diesel fuel in the combustion chamber, leading to increases in the temperature and pressure, and facilitating a more complete combustion of fuel during the main injection. When the energy ratio of ammonia is low, the diesel fuel pre-injection ignites and burns the ammonia mixture before the main injection. The ignition is controlled by the kinetics of the chemical reaction at this moment. At a high ammonia energy ratio, the proportion of premixed ammonia is high, and the latent heat of ammonia vaporization strengthens the effect of lowering the cylinder temperature, which reduces the ignition effect of the pre-injected diesel fuel, and the ignition is determined by the amount of diesel fuel injected into the cylinder at the moment of the main injection. At an ammonia energy ratio exceeding 70%, the in-cylinder pressure is lower [24] with the pre-injection strategy compared to the single-injection strategy. This is primarily due to the decrease in the main injection diesel fuel and the decentralized injection timing, resulting in a lower concentration of diesel fuel in the cylinder and a thinner ignition fuel mixture. Consequently, there is a weakened ignition intensity, slower combustion, and a delayed combustion stage, leading to a reduced in-cylinder pressure.

Figure 4 illustrates the changes in the combustion phase of an ADDF engine with a varying ammonia energy ratio under the single-injection and pre-injection diesel injection strategies. The pre-injection delay time is calculated from the diesel pre-injection timing. As the ammonia energy ratio increases, the stagnation period gradually lengthens, and the in-cylinder combustion duration shows a tendency to become shorter and then longer. The characteristics of ammonia used in engines, such as difficult ignition, slow combustion, and difficulty in spontaneous combustion [25], inhibit the spontaneous combustion of diesel fuel, resulting in a delay in the ignition start point and a delay in the center of gravity of combustion. The combustion duration compared to the ammonia ratio of 0% (i.e., pure diesel mode) is significantly reduced, the liquid ammonia in the diesel fuel injection in advance of the injection and the formation of a homogeneous mixture under the action of the air flow in the cylinder, and the local high-temperature combustion of diesel fuel gradually diffuse and ignite the NH₃ mixture in the cylinder so that the proportion of cylinder-premixed combustion is increased [26], the combustion rate is accelerated, the exothermic process is more centralized, and the combustion duration is shortened. At a high ammonia energy ratio, lengthening the stagnation period and reducing the diesel fuel duty cycle increases the diffusion time of the ignited diesel fuel in the cylinder, decreases the local equivalence ratio, delays the time it takes to reach the ignition limit, and increases the delay effect of the stagnation period. As the energy ratio of ammonia increases, the diesel fuel ratio decreases, and the higher latent heat of the vaporization of ammonia lowers the in-cylinder temperature, prolonging the time it takes to reach the minimum combustion temperature, while the low combustion temperature of ammonia further inhibits the combustion process of the mixture inside the cylinder, leading to an increase in the combustion duration under the high ammonia energy ratio. According to the findings from Figure 4b, the combustion duration is reduced when pre-injection is added under the same ammonia energy ratio. Specifically, the combustion duration is shortened by approximately 17.35% when the ammonia energy ratio is at 70%. The pre-injection of diesel fuel increases the reactivity of the NH₃ mixture, increases the in-cylinder temperature during the compression stroke, improves the evaporative atomization of the main injection of diesel fuel, and promotes the complete combustion of ammonia, which leads to a shorter combustion duration.

Figure 4c illustrates the changes in the single-injection and pre-injection strategies on the indicated thermal efficiency (ITE) of ADDF combustion modes. As the ammonia energy ratio increases, the ITE of the ADDF first increases and then decreases, ADDF combustion significantly improves the engine ITE, and the addition of the pre-injection of primed diesel results in a significant increase in the ITE. At a 70% ammonia energy ratio, the increase in the ITE is approximately 22.86% for single-injection diesel and 21.25% for pre-injection in comparison to the diesel-only fuel. The peak rate of heat release is significantly higher in the ADDF combustion mode compared to the diesel-only mode. According to the previous analysis, implementing a pre-injection strategy resulted in a more advanced and concentrated exothermic process and biased the combustion process towards fixed-volume combustion. This led to an improvement in the ITE compared to the single-injection strategy, with approximately a 2% increase at a 70% ammonia energy ratio. When the energy ratio of ammonia exceeds 70%, a small quantity of diesel is insufficient to provide the necessary ignition energy for the mixture of NH₃ and air, resulting in prolonged heat release and the deterioration of combustion, which reduces the ITE. Figure 4c displays that the ITE of the pre-injection strategy is lower than the ITE of the single-injection strategy for ammonia energy ratios of 20% and 40%. Additionally, the combustion start point is advanced in pre-injection, and most of the fuel takes part in combustion before the top dead center. This leads to an increase in the negative work of combustion, which, in turn, decreases the ITE.

Table 5. (a) In-cylinder temperature field of ADDF engine with single-injection strategy under various ammonia ratios. (b) In-cylinder temperature field of ADDF engine with pre-injection strategy under various ammonia ratios.

(a)
	Crank Angle	0° CA ATDC	4° CA ATDC	8° CA ATDC	12° CA ATDC	20° CA ATDC
Ammonia Energy Ratio
0%
20%
40%
60%
65%
70%
75%
(b)
	Crank Angle	0° CA ATDC	4° CA ATDC	8° CA ATDC	12° CA ATDC	20° CA ATDC
Ammonia Energy Ratio
0%
20%
40%
60%
65%
70%
75%

presents the change in the ammonia energy ratio on the in-cylinder temperature field of the ADDF engine with the single-injection and pre-injection strategies for diesel. From the table, it can be seen that the time of appearance of the high-temperature region is continuously delayed with the increase in the ammonia energy ratio, the addition of ammonia has an inhibiting effect on the combustion of diesel fuel, and the period of ignition delay is continuously prolonged. The high-temperature area initially appears in the pit area near the intake side, which is the first to attain ignition conditions, primarily because the denser concentration of ignited diesel fuel occurs near the intake side compared to the rest of the combustion chamber upon approaching the top dead center, and because combustion typically begins at the end of the diesel fuel spray [27] and then gradually diffuses throughout the cylinder. As the energy ratio of ammonia increases in the dual-fuel mode, the high-temperature region under the same crankshaft angle gradually decreases. In comparison to the pure diesel mode, the temperature in the unburned region is much lower. This is mainly due to the lower combustion temperature of ammonia, the slow propagation of the flame, and the delay of the combustion stage. Consequently, a significant amount of unburned ammonia exists in the cylinder. In Table 5(b), the pre-injection strategy results in a forward shift of the combustion starting point at the same ammonia energy ratio, as well as an increase in the in-cylinder temperature at the same crankshaft angle, when compared to the single-injection strategy. The pre-injection of diesel fuel not only improves the ignition performance of the NH₃ mixture, increases the cylinder temperature, and promotes the flame propagation in the cylinder, but also makes the main injection of diesel fuel reach the ignition state earlier, so that the cylinder combustion is advanced.

3.1.2. Effects of Ammonia Energy Ratios on Emission Characteristics

Figure 5 illustrates the impact of different energy ratios of ammonia on the emissions of a dual-fuel engine. According to Figure 5(a1,a2), the emissions of soot consistently decrease with an increase in the ammonia energy ratio. The maximum soot emission occurs when only diesel is used. However, a 70% ammonia energy ratio reduces the emission peaks by 98.13% and 99.6%. The soot emissions decline as a result of the reduction in diesel injection, the carbon concentration within the fuel cylinder, and a drop in the combustion temperature within the cylinder. According to Figure 5(a1,a2), the emission of NO_x increases and then decreases as the ammonia energy ratio increases. The peak emission occurs at 60% and 20% ammonia energy ratios. Furthermore, when compared to it, NO_x emission decreases by 38.73% and 37.63% at a 70% ammonia energy ratio. Compared to the mode of diesel-only fuel, when the ammonia energy ratio is increased, there is a rise in the nitrogen content of the fuel inside the cylinder. This, in turn, leads to an increase in the emissions of NO_x. As the ammonia energy ratio increases in the dual-fuel mode, the combustion starting point and center of gravity are delayed and move away from the TDC. The majority of exothermic combustion processes occur after reaching the TDC, which leads to drops in the cylinder pressure and temperature. The amino group produced during the ammonia reaction has the effect of denitrogenation, which can consume NO_x, and all of this leads to a reduction in NO_x emissions. Pre-injection raises the in-cylinder temperature compared to the single-injection strategy, resulting in increased NO_x emissions at the same ammonia energy ratio.

In Figure 5(a1,a2), the HC emissions decrease and then slightly increase, and the CO emissions gradually decrease as the ammonia energy ratio increases. As the energy ratio of ammonia increases, the cylinder receives less diesel fuel injection, which leads to reduced emissions of HC and CO. The rise in the temperature within the combustion chamber during the combustion process enhances fuel combustion and minimizes HC emissions. Conversely, when the ammonia energy ratio grows, the in-cylinder temperature drops, inhibiting complete fuel combustion and causing an increase in HC emissions. In the pre-injection strategy, the delayed center of gravity of combustion and premix accumulation resulted in a noteworthy reduction in the CO emissions when compared to the single direct-injection strategy. At an ammonia energy ratio of 70%, the CO emissions were reduced by 97.13%.

In Figure 5(b1,b2), there is a significant decrease in the CO₂ emissions as the ammonia energy ratio increases. This decrease is due to a proportional reduction in CO₂ resulting from the decrease in carbon-containing diesel fuel. The N₂O emissions first increase, then decrease, and then increase after a 70% ammonia energy ratio. The N₂O generated by the oxidation of NH₃ decomposes at temperatures above 1073–1273 K [28]. The formation and breakdown of N₂O correlates with the combustion rate of ammonia and the temperature within the cylinder [29]. As the ratio of ammonia energy increases, the nitrogen content in the fuel also increases, leading to a notable surge in N₂O emissions in comparison to the pure diesel mode. In the dual-fuel mode, the combustion temperature within the cylinder decreases with an increase in the ammonia energy ratio due to the slow flame propagation and low flame temperature of ammonia [25]. As a result, the emissions of N₂O decrease. However, at a 75% ammonia energy ratio, the N₂O emissions increase due to excessive hysteresis during the combustion phase.

N₂O is the main product of ammonia combustion, and its effect on greenhouse gases is considered to be 300 times that of CO₂ [30]. The emissions of GHG (including N₂O and CO₂; the equivalent calculation formula is GHG = CO₂ + N₂O × 300) under different ammonia energy ratios are shown in Figure 5(b1,b2), from which it is obvious that GHG emissions gradually decrease as the ammonia energy ratio increases. GHG emissions are the lowest when a 70% energy ratio is used. Compared with the pure diesel mode, the ADDF mode has an obvious emission reduction effect, and GHG emissions are reduced by 36.75% under the single-injection strategy. The pre-injection strategy is reduced by 40.9%.

In Figure 5(b1,b2), the emission of unburned NH₃ increases initially before decreasing as the ammonia energy ratio rises. NH₃ emission reaches its minimum at a 70% ammonia energy ratio, and the pre-injection strategy significantly reduces the NH₃ emission by 9.3% at a 70% ammonia energy ratio. As the energy ratio of ammonia increases and more diesel is replaced by it, there is a noteworthy rise in unburned NH₃ emissions compared to the model using only diesel. As the energy ratio of ammonia increases in the dual-fuel mode, the combustion duration increases, resulting in a decrease in unburned NH₃ in the cylinder and emissions. Reducing diesel fuel volumes and delaying combustion stages lead to the enhanced reduction in unburned NH₃ emissions at a 75% energy ratio of ammonia.

Table 6. (a) Mass fractions of N₂O and NH₃ of ADDF engine with single-injection strategy under high ammonia ratio. (b) Mass fractions of N₂O and NH₃ of ADDF engine with pre-injection strategy under high ammonia ratio.

(a)
			Crank Angle	0° CA ATDC	4° CA ATDC	8° CA ATDC	12° CA ATDC	20° CA ATDC
Ammonia Energy Ratio
60%		N2O
		NH3
65%		N2O
		NH3
70%		N2O
		NH3
75%		N2O
		NH3
(b)
			Crank Angle	0° CA ATDC	4° CA ATDC	8° CA ATDC	12° CA ATDC	20° CA ATDC
Ammonia Energy Ratio
60%	N2O
	NH3
65%	N2O
	NH3
70%	N2O
	NH3
75%	N2O
	NH3

illustrates the variation of the concentration fields of N₂O and NH₃ under a high ammonia energy ratio in the cylinder. N₂O is mainly distributed along the flame front of NH₃, and the concentration of N₂O is the highest at the boundary of unburned NH₃. Compared with the temperature field in Table 5, it is found that N₂O is mainly located in the low-temperature zone in the cylinder, and it almost does not exist in the high-temperature zone [10]. The variation of the NH₃ concentration field in the cylinder is basically consistent with that of the temperature field. The main reason for the production of N₂O is due to the decomposition of NH₃ [31]. In the early stage of engine compression ignition combustion, the NH₃ combustion reaction in the cylinder quickly generates N₂O, causing a rapid increase in N₂O. With the rapid increase in the cylinder temperature to the N₂O decomposition temperature or higher, N₂O will be thermally decomposed so that its generation rate is reduced, and at this time, the piston is away from the TDC of the process, the temperature inside the cylinder decreases, the combustion rate slows down, the NH₃ reaction rate slows down so that the thermal decomposition of N₂O dominates the process, and the concentration of N₂O is reduced.

3.2. Effects of Injection Timing on Combustion and Emission Characteristics

3.2.1. Effects of Injection Timing on Combustion Characteristics

Figure 6 displays the impact of the ignition diesel injection timing on the in-cylinder pressure for the ADDF engine under a 70% ammonia energy ratio and ignition diesel pre-injection strategy conditions. In Figure 6a, when the main injection timing of diesel fuel is advanced, the peak in-cylinder pressure shows an upward trend, and its phase corresponding to the peak gradually approaches the top dead center (TDC). Advancing the main injection timing from −4° CA ATDC to −14° CA ATDC results in a 25.82% amplification in the peak cylinder pressure, rising from 10.03 MPa to 12.62 Mpa. This phenomenon stems from the advance of the main injection timing, resulting in an increased mixing time for the cylinder diesel fuel and the NH₃-air mixture, allowing for a complete fuel mixture in the cylinder. This leads to a dominance of premixed combustion, which causes the cylinder temperature to rise, resulting in a further increase in the cylinder pressure. Meanwhile, with the advance of the main injection timing, the fresh cylinder mixture reaches the ignition conditions to shorten the time, resulting in the advance of the combustion phase. In Figure 6b, a gradual decrease in the peak cylinder pressure is shown as the diesel pre-injection timing is advanced, and the peak phase gradually shifts away from the TDC. The timing of the pre-injection is advanced from −50° CA ATDC to −18° CA ATDC, resulting in a rise in the peak cylinder pressure from 10.55 Mpa to 11.41 Mpa by 8.15%. Advancing the pre-injection timing extends the ignition delay period, which allows the diesel fuel and ammonia mixture to form a homogenous premixed gas. Consequently, the reactant activity is reduced, the ignition energy and ignition strength are decreased, the combustion initiation is delayed, and ultimately, there is a reduction in the cylinder pressure. But during the pre-injection timing between −34° CA ATDC and −50° CA ATDC, there is a minimal change in the cylinder pressure. This occurs because the pre-injection timing of diesel fuel happens too early, leading to the formation of a too-thin premix concentration. Under such circumstances, the cylinder temperature and pressure are low, resulting in the diesel fuel pre-injection being unable to ignite the NH₃ mixture. As such, the mixture ignites mainly due to the main injection of diesel fuel [32]. Therefore, the alteration in the cylinder pressure curve remains insignificant. The pre-injection timings at −26° CA ATDC and −18° CA ATDC result in a significant increase in the in-cylinder pressure. The reductions in the timing interval between the diesel pre-injection and main-injection timings play the most important roles in this outcome. This results in the main injection of diesel fuel being close to the formation of the pre-injection of diesel fuel in the local high-temperature region. This effectively improves the cylinder ignition energy and ignition intensity, promoting the main injection of diesel fuel ignition for the NH₃ fuel mixture. Additionally, while the pre-injection timing is later, the in-cylinder temperature and pressure are high, and the stagnation period is shortened, and as a result, there is a significant increase in the in-cylinder pressure.

Figure 7 illustrates the effects of the changes in the pre-injection strategy on the in-cylinder combustion phases of the ADDF engine, considering the priming diesel main injection and pre-injection timings. Figure 7a shows that as the diesel main injection timing is delayed, the ignition delay period gradually decreases and is minimized at −6° CA ATDC. Additionally, the combustion duration period initially decreases and then increases, with the shortest duration observed at −8° CA ATDC. Delaying the timing of the main injection of diesel fuel reduces the diffusion time of the diesel spray, improving the local equivalent ratio. When the moment of injection is close to the TDC, the temperature within the cylinder is higher, which is conducive to the atomization of the main injection of diesel fuel. This, in turn, allows the diesel fuel within the cylinder to reach ignition conditions more quickly, resulting in an earlier exothermic combustion process. The analysis shown in Figure 7b demonstrates that as the timing of the diesel pre-injection is postponed, the ignition delay period initially rises before declining, and the combustion duration is gradually lengthened. Delaying the pre-injection timing diminishes the atomization and vaporization diffusion duration of the pre-injected diesel inside the cylinder, increases the local concentrated mixing zone near the pre-injection jet, and makes the combustion start time gradually advance. As shown in

Table 7. Comparison of in-cylinder engine temperatures and local equivalence ratios at two pre-injection timings.

Pre-Injection Timing (° CA ATDC)	Temperature		Equivalent Ratio
	After Pre-Injection 2° CA	−6° CA ATDC	−6° CA ATDC
−26
−42

, when the pre-injection timing is set at a −42° CA ATDC (After Top Dead Center), the in-cylinder temperature at a 2° CA (Crank Angle) after pre-injection is significantly low. During this time, the pre-injected diesel undergoes a relatively long evaporation process and effectively mixes with the NH3-air mixture in the cylinder during the ignition delay period. At a −6° CA ATDC, although the temperature of in-cylinder combustion is sufficiently high, the premixed fuel–air mixture is relatively thin. At this time, the relatively small mixture equivalence in the cylinder leads to a delay in the combustion start point. When the pre-injection timing is set at a −26° CA ATDC, the cylinder temperature is higher at a 2° CA after pre-injection. This leads to a shorter ignition delay period, an increased local equivalence ratio, an increased promotion of the main injection ignition, an advanced moment of ignition, a reduced formation of premix by the main injection, and an increased combustion duration period. The earlier pre-injection timing allows for the pre-injected diesel to mix more evenly with the ammonia–air mixture, while the NH₃ mixture mixed with diesel has a higher flame propagation rate, so the combustion duration is relatively shorter.

3.2.2. Effects of Injection Timing on Emission Characteristics

Figure 8 illustrates that the emission characteristics of CO₂, N₂O, GHG, and NH₃ of a dual-fuel engine vary with the main injection timing and the pre-injection timing under the pre-injection strategy. Figure 8a displays a pattern where the CO₂ and N₂O emissions decrease and then increase with the delay of the main injection timing. The CO₂ emissions reach the lowest point at the injection timing of a −6° CA ATDC, and the N₂O emissions reach the lowest point at the injection timing of −12° CA ATDC. Figure 8b shows that both the CO₂ and N₂O emissions show a downward trend as the pre-injection timing advances, with the N₂O emissions being the lowest when the pre-injection timing reaches −42° CA ATDC. This phenomenon is attributed to the advancement of the main injection timing and pre-injection timing, which contribute to the increase in the cylinder temperature, thereby facilitating the decomposition of N₂O and thus reducing its emission. From Figure 8, it can be seen that the trend of the GHG emissions at the injection timing is consistent with the trend of N₂O, so N₂O dominates the GHG emissions at a fixed ammonia energy ratio. The minimum emission level was observed during the main injection timing at −12° CA ATDC and during the pre-injection timing at −50° CA ATDC, resulting in 9.52% and 16.77% reductions in emissions compared to the peak levels. By adjusting the injection timing appropriately, the GHG emissions can be reduced by 69.34% under the pre-injection strategy when compared to the pure diesel mode.

In Figure 8, the emission of unburned NH₃ shows a slight decrease at first and then an increase as the injection timing of diesel is delayed. The emission is at its lowest when the main injection timing is −10° CA ATDC. The delayed main injection timing delays the combustion stage, causing part of the NH₃ in the cylinder not to burn in time. Consequently, unburned NH₃ emissions increase significantly at the main injection timing of −4° CA ATDC. The lowest NH₃ emission was found when the pre-injection timing was −42° CA ATDC. Compared to Figure 5(b2), the emission of NH₃ at −26° CA ATDC is significantly decreased from 2.27 g/(kW∙h) to 8.81 × 10⁻⁵ g/(kW∙h), mainly due to the injection timing of NH₃. Advancing the timing of NH₃ injection improves the quality of the mixture inside the cylinder and the combustion quality.

4. Conclusions

In this study, CFD was used to simulate the combustion of an ADDF cylinder high-pressure direct-injection engine, and the combustion and emission characteristics of the ADDF under different ammonia energy ratios (0%, 20%, 40%, 60%, 65%, 70%, 75%, and 80%) and diesel injection timings (pre-injection timing and main injection timing) were discussed. The main conclusions are summarized as follows:

(1)

In the dual-fuel mode, the peak cylinder pressure decreased, and the combustion phase was delayed by increasing the ammonia energy ratio; the NO_x emissions first increased and then decreased; the CO and CO₂ emissions decreased; and the N₂O emissions decreased but increased when the ammonia energy ratio exceeded 70%.

(2)

Compared with the single-injection strategy, the peak value of the cylinder pressure under the pre-injection strategy is increased, the combustion stage is advanced, the ITE is increased, there is a 2% improvement at a 70% energy ratio with ammonia, and the soot and CO and NH₃ emissions are decreased, but NO_x emissions are increased.

(3)

The ADDF engine has a better overall performance at an ammonia energy ratio of 70% compared with the pure diesel mode, the increase in the ITE is approximately 22.86% for the single-injection strategy and 21.25% for the pre-injection strategy for diesel. At present, the ADDF engine has an obvious GHG emission reduction effect. The GHG emissions (including CO₂ and N₂O) are reduced by 36.75% under the single-injection strategy and by 40.9% under the pre-injection strategy. The ADDF engine has the lowest NH₃ emissions, with a 9.3% reduction compared to the single-injection strategy.

(4)

As the main injection timing of diesel fuel advances, the peak cylinder pressure increases, and the combustion stage is advanced. As the pre-injection timing advances, the peak cylinder pressure decreases, causing a slight backward shift in the combustion stage. When the ammonia injection timing is −150° CA ATDC, the main injection timing of pilot diesel is −12° CA ATDC, and the pre-injection timing is −50° CA ATDC; the NH₃ emissions are significantly reduced; and the GHG emissions are reduced by 69.34% compared to the diesel-only mode with pre-injection.

In conclusion, this study discusses and analyzes the ammonia duty cycle and fuel injection timing related to the ADDF engine and suggests that a 70% ammonia duty cycle is feasible in the studied engine model and achieves a high indicated thermal efficiency and low emissions, and the subsequent study of the injection timing is also based on this duty cycle. The above studies were carried out based on CFD simulations, but they can be validated later using engine bench tests, after which these findings can be an inspiration and contribution to academic research and have practical applications in related fields, as well as serve as a guideline for the preparation of NH₃ fossil diesel blends.