Research on Combustion Characteristics of Ammonia/N-Heptane Dual-Fuel Marine Compression Ignition Direct-Injection Engine

Abstract

To address the decarbonization requirements of the shipping industry, this study establishes an in-cylinder combustion simulation model for a medium–high speed four-stroke ammonia-fueled marine engine based on the CONVERGE v3.0 platform. A diesel combustion model was first developed and validated against experimental data. Building on this validated model, an ammonia/n-heptane dual-fuel combustion model was further developed by coupling a chemical kinetic mechanism for ammonia/n-heptane. To overcome the challenge of igniting pure ammonia, a combustion strategy employing intake port injection of n-heptane and direct in-cylinder injection of ammonia fuel was adopted, leveraging thermal compression ignition. The results indicate that under initial cylinder conditions of 1 bar and 350 K, misfire occurs when the ammonia energy proportion (AEP) reaches 70%, preventing stable ignition and combustion of ammonia. Based on an analysis of intake boundary conditions, the influence of intake supercharging coupled with intake heating on ammonia combustion characteristics was investigated. As the AEP increases further, the combustion of n-heptane deteriorates significantly. At a 90% AEP, the combustion efficiency of n-heptane is approximately 67% at an initial temperature of 350 K but drops to about 28% at 400 K.

1. Introduction

Thanks to its excellent thermal efficiency, high reliability, long service life, and direct compatibility with propellers, piston internal combustion engines (ICEs) currently hold a core dominant position in the international maritime main propulsion system market, a pattern expected to persist in the foreseeable future. The International Maritime Organization (IMO) has set a landmark target of achieving net-zero greenhouse gas (GHG) emissions by 2050 [1], which is more ambitious than its initial strategy in 2018; meanwhile, the European Union Emissions Trading System (EU ETS) has been extended to the shipping sector since 2025 [2].

Against this policy backdrop, the adoption of alternative fuels in internal combustion engines is widely recognized as the most practical and effective pathway to achieve the aforementioned emission reduction goals [3]. Fossil fuels (primarily petroleum) are the main source of carbon emissions. As GHG emissions (especially carbon dioxide) continue to rise and the greenhouse effect intensifies, carbon reduction has become a global consensus, leading to widespread attention on the application of clean alternative fuels such as hydrogen and ammonia. Among these, carbon-free ammonia, with advantages including convenient storage, feasible large-scale promotion, strong sustainability, and significant cost-effectiveness [4,5], has combustion products mainly consisting of nitrogen oxides, nitrogen, and water. It also boasts a high-octane number, strong renewability, and high energy density. Additionally, ammonia is easy to liquefy at room temperature, featuring high storage density and convenient transportation and utilization—a characteristic that distinguishes it from hydrogen—making it one of the most feasible alternative fuel options for the shipping industry.

The application potential of ammonia as an engine fuel has attracted extensive attention from researchers. High-pressure direct injection (HPDI) has been initially selected as the technical direction by MAN Energy Solutions [6], the leading enterprise in the marine main engine market, due to its effectiveness in reducing ammonia slip, improving charging efficiency, and offering outstanding control flexibility. In recent years, this mode has also gained focused attention in academia, with research covering areas such as fundamental theories of spray combustion [7,8,9] and practical engine applications [10,11]. Typically, in addition to the HPDI ammonia injector, the cylinder head is also equipped with a reserved diesel injector, which is used to inject highly reactive pilot diesel near top dead center (TDC) to initiate the combustion process [12].

However, ammonia faces inherent technical bottlenecks in engine applications, specifically manifested as high ignition energy requirements, a narrow flammability limit, low laminar flame speed, and unstable combustion processes [13]. Consequently, traditional spark ignition (SI) and compression ignition (CI) engines cannot operate solely on ammonia as a single fuel: SI engines require an increased compression ratio or intake air temperature to achieve reliable ignition [14], yet even so, pure ammonia combustion in SI engines still underperforms compared to conventional gasoline SI engines, with issues such as decreased fuel efficiency and intensified incomplete combustion. Hoseinpour M et al. [15] present a numerical study using CFD simulations to explore how varying inlet charge temperatures and EGR rates influence the combustion and emissions of an ammonia/diesel dual-fuel engine under high-load conditions, identifying optimal conditions for significant CO₂ and NO₂ reductions and peak NH3 reduction. Cameretti et al. [16] performed CFD simulations using ANSYS Forte^® (2024 R2) and diesel injection strategies in an ammonia–diesel dual-fuel marine engine to maintain engine performance and limit ammonia exhaust losses, identifying an optimal split injection strategy that also quantifies NO_x and N₂O emissions. Kakoee A et al. [17] show promise for marine engines due to their potential for low NO_x emissions and high thermal efficiency, yet their performance is sensitive to factors like injection timing; this study employs computational fluid dynamics (CFD) to examine how the start of ignition (SOI) impacts marine RCCI engine performance. Palomba M et al. [18] use 3D-CFD simulations to explore combustion and emissions in a heavy-duty ammonia-fueled engine with a pre-chamber spark ignition system, finding optimal performance with a methane-fueled pre-chamber, 80% NH3–20% CH4 main chamber blend, and 38 BTDC spark timing. To optimize the combustion performance of ammonia engines, researchers have developed various technical solutions [19,20,21], among which auxiliary combustion technology has proven to be the most effective. By blending ammonia with high-reactivity fuels like hydrogen [22], diesel [23,24], gasoline [25,26], methanol [27,28,29,30,31], or methane [32], the problem of slow ammonia flame propagation can be resolved, thereby improving combustion efficiency. For CI engines, high-reactivity auxiliary fuels such as diesel or dimethyl ether are required to trigger high-temperature ignition, an operating mode defined as dual-fuel (DF) compression ignition [33]. Due to the typically higher compression ratio of CI engines compared to SI engines, CI engines under the dual-fuel compression ignition mode offer higher thermal efficiency and lower incomplete combustion losses, making them the preferred power unit for heavy-duty equipment such as ships, trucks, and generators.

In recent years, researchers have gradually focused on the operating characteristics of marine direct-injection compression ignition ammonia-fueled engines. To deeply analyze the combustion mechanism of diesel/ammonia, Pan [34] adopted a combination of numerical simulation and experimental testing to explore the oxidation reaction laws of ammonia/n-heptane. The results showed that under moderate temperature conditions, there is a mutual promotion effect between their oxidation reactions. Early research on the ammonia dual-fuel compression ignition (DF CI) mode was pioneered by Pearsall and Garabedian (1967) [35]. By analyzing the impact of different pilot fuels on engine power output and fuel consumption, they found that high cetane number fuels can effectively optimize engine performance; the optimal fuel economy of the engine is achieved when the total equivalence ratio (φ) is in the range of 0.7–0.75; and the maximum power output is obtained when the total equivalence ratio is approximately 0.9. Reiter and Kong [36] systematically investigated the influence of the pilot fuel energy fraction on combustion characteristics and emission performance in a four-cylinder ammonia DF CI engine. The results indicated that achieving an ideal fuel economy requires maintaining a pilot fuel energy fraction of 40–60%, and nitrogen oxide emissions increase significantly when this fraction exceeds 50%.

Scholars have conducted extensive research on ammonia fuel injection strategies. Niki et al. [37] pointed out that ammonia injection in the intake port leads to excessive ammonia emissions; although pilot injection or post-injection technology can reduce ammonia emissions, it is accompanied by an additional increase in nitrous oxide (N₂O) emissions. Jin et al. [38] recently proposed a split injection strategy, which successfully achieved effective control of N₂O emissions under the operating condition of 50% ammonia energy fraction. However, it should be noted that N₂O has a global warming potential approximately 300 times that of carbon dioxide, and even with this strategy, the net reduction in greenhouse gas emissions is only 14.2% higher than that of pure diesel combustion. Therefore, it is necessary to equip aftertreatment systems and advanced combustion strategies to further reduce the total greenhouse gas emissions. Liu [39] et al. proposed the concept of Reactivity-Controlled Turbulent Jet Ignition (RCTJI) for ammonia engines, in which a newly designed air-assisted pre-chamber system with scavenging and hydrogen injection is employed.

In the research on marine direct-injection ammonia engines, Park et al. [40] supported ammonia ignition with pilot fuel through control strategies and optimized the injection timing and quantity of the pilot fuel based on operating conditions. The study adopted high-compression-ratio pistons and experimentally observed their impacts on combustion stability and exhaust emissions in a marine high-pressure ammonia direct-injection dual-fuel engine.

For marine compression ignition ammonia engines, Liu et al. [41] conducted extensive computational studies to tap the application potential of ammonia in internal combustion engines, focusing on investigating the effects of two advanced combustion technologies—pre-chamber combustion (PCC) and dual-fuel compression ignition (DF-CI)—on heavy-duty ammonia-fueled engines.

Dong et al. [42] demonstrated that the ignition chamber injection ignition system outperforms the spark plug ignition system in enhancing ammonia combustion and shortening the combustion duration. The study evaluated the combustion performance of ammonia under different equivalence ratios (1.0 and 0.8) and compared the influences of different gasoline energy percentages (2.5%, 2.0%, 1.5%, and 1.0%) on its combustion characteristics.

In numerical simulations of combustion processes, for ammonia/diesel dual-fuel engines, Ou et al. [43] developed a simplified ammonia/n-heptane chemical kinetic mechanism (with n-heptane as a single-component surrogate for diesel). This mechanism includes detailed ammonia/n-heptane chemical reactions and is applicable to multi-dimensional computational simulations in scenarios such as construction machinery and ships. Huang et al. [44] investigated ammonia/n-heptane dual-fuel combustion under high pressure via direct numerical simulation (DNS) to clarify the effect of ammonia addition on n-heptane spray combustion behavior; their findings provide important insights into the micro-mechanism of mixed combustion between ammonia and high-reactivity fuels. Meng et al. [45] also used three-dimensional direct numerical simulation to explore the process of n-heptane flame igniting ammonia spray, subsequent flame development, and interactions between diesel/ammonia flames under high pressure, with a particular focus on diesel/ammonia direct dual-fuel stratified (DDFS) engine conditions. These DNS studies are crucial for understanding combustion complexity and optimizing dual-fuel engines. Zhou et al. [46] conducted in-depth research on ignition characteristics, combustion modes, and NO/N₂O emissions of ammonia/n-heptane combustion under Reactivity-Controlled Compression Ignition (RCCI) engine conditions, aiming to address challenges such as low flame speed and high nitrogen oxide emissions when ammonia is used as a marine engine fuel.

Currently, ammonia injection strategies as independent variables have not been extensively studied in experiments, and experimental data are urgently needed for support. Compared with other fuels, ammonia has a higher auto-ignition temperature and requires significantly more ignition energy, indicating strong resistance to auto-ignition and ignition. This makes it difficult to achieve stable ignition and combustion in both compression ignition and spark ignition engines. Due to the slow flame propagation speed of ammonia combustion, the traditional ammonia–diesel dual-fuel RCCI combustion method (ammonia injected into the intake port) has certain limitations. In this study, an ammonia/n-heptane dual-fuel combustion simulation model was constructed based on the CONVERGE simulation platform. By coupling the fuel chemical reaction kinetic model, the combustion characteristics of ammonia under thermal atmosphere conditions were studied in depth. The combustion characteristics under different ammonia substitution rate conditions were analyzed, and the influence of the compression ratio, intake boundary conditions, and ammonia injection strategies on in-cylinder combustion characteristics and nitrogen oxide emissions was explored. This study adopts in-cylinder direct injection of ammonia, which can achieve higher combustion efficiency and thermal efficiency. The analysis of this research can provide a basis and guidance for further exploration of ammonia fuel combustion, with important theoretical value and practical application significance.

2. Numerical Model and Validation

2.1. Engine Operating Parameters

The experimental study utilized a MAN L23/30H four-stroke compression ignition engine, designed as a marine propulsion unit with enlarged cylinder dimensions. According to the classification criteria for marine engine rotational speeds (low speed: <300 rpm, medium and high speed: 300–2000 rpm), the MAN L23/30H model (1500 rpm) used in this study falls into the medium- and high-speed category. The principal technical specifications for this L23/30H power source are provided in

Table 1. Main technical specifications of the MAN L23/30H single-cylinder diesel engine.

Parameter	Value
Cylinder diameter	105 mm
Piston stroke	125 mm
Connecting rod length	210 mm
Engine speed	1500 r/min
Indicated Mean Effective Pressure (IMEP)	12.6 bar
Maximum Combustion Pressure (MCP)	150 bar
Compression ratio	16
Single-engine rated output	175 kW
Intake Valve Closing	−133 °CA
Exhaust Valve Closing	−14 °CA
Pilot fuel injection timing	−7 °CA

. This engine integrates three essential technologies: (1) a common-rail direct-injection system, (2) an electro-hydraulic fuel supply mechanism, and (3) a durable fuel management setup. Its turbocharging arrangement comprises a single-stage ABB A165-L turbocharger (ABB, Baden, Switzerland), coupled with a KLQ50H intercooler (produced by ZC Diesel, Zibo, China) that employs single-stage water cooling. The experiment adopted a baseline diesel operating condition (turbocharged, with an intake pressure of 2 bar and a rotational speed of 1500 rpm). The simulated operating conditions were divided into two categories, naturally aspirated (1 bar) and turbocharged (2 bar), corresponding to different intake boundary conditions, respectively. The MAN L23/30H features a bowl-type combustion chamber configuration, with the combustion chamber contour depicted in Figure 1b. The engine’s external design and physical configuration are depicted in Figure 1.

2.2. Model Establishment

This research examines a heated-charge compression ignition mode utilizing direct in-cylinder ammonia injection combined with port-injected n-heptane. Starting from a validated baseline diesel engine model, an ammonia-fueled combustion model was developed by integrating a detailed ammonia/n-heptane chemical kinetics mechanism. Core engine geometry parameters—including bore, stroke, and compression ratio—were retained from the original model, with the primary modification being the fuel system.

The chemical kinetic mechanism employed is the ammonia/n-heptane blend mechanism published by Xu et al. [47]. This mechanism was constructed by integrating and refining several established sub-mechanisms: a skeletal n-heptane model [48], a machine learning-optimized ammonia mechanism [49], and foundational H₂/O₂ [50] and NO_x [51,52,53] reaction sets. The final combined mechanism comprises 69 species and 389 reactions. Validation by Xu et al. [47] confirms its accuracy in predicting ignition and combustion for ammonia and ammonia/n-heptane blends with manageable computational cost.

For the simulations, the total fuel energy input was held constant. The required masses of ammonia and n-heptane were calculated accordingly. Ammonia was directly injected into the cylinder, while n-heptane was premixed with the intake charge. This configuration creates a dual-fuel system where the high-reactivity n-heptane mixture ignites first, generating a reactive thermal atmosphere that facilitates the compression ignition and subsequent diffusion combustion of the directly injected ammonia. This approach is termed Ammonia Direct-Injection Thermal Atmosphere Compression Ignition.

N-heptane was chosen as the port-injected fuel due to its high cetane number (promoting reliable ignition), well-defined chemical properties, and lower viscosity—which favors evaporation and mixing in a premixed charge. To reduce computational expense, a 45-degree sector model (1/8 of the full geometry) was adopted, leveraging the axisymmetric combustion chamber and the even distribution of eight fuel injector nozzles. The final simulation model is illustrated in Figure 2.

For fuel quantity determination, the total energy input was first calculated based on the original diesel fuel mass (lower heating value: 42.8 MJ/kg). The respective masses of ammonia (LHV: 18.6 MJ/kg) and n-heptane (LHV: 44.9 MJ/kg) were then derived for each operating condition according to the specified ammonia energy proportion (AEP).

In the Direct-Injection Dual-Fuel (DIDF) configuration, ammonia is injected directly into the cylinder. To compensate for its lower energy density and maintain an injection duration comparable to the baseline diesel case, the injector nozzle diameter was increased to 0.24 mm while keeping the number of holes unchanged at eight. The corresponding simulation model is presented in Figure 3.

2.3. Simulation Submodel Settings

Turbulence parameters at intake valve closure were calibrated based on engine geometry. In-cylinder turbulent flow was simulated using the Reynolds-Averaged Navier–Stokes (RANS) equations with the RNG k-ε model [54]. Droplet breakup was described by the KH-RT model, with its constants set according to Shi et al. (see

Table 2. Settings of parameters of the KH-RT model.

Submodel	Parameter Values	Diesel Spray Model	Ammonia Spray Model
KH-model	Size constant B0	0.61	0.61
	Velocity constant C1	0.188	0.188
	Breakup time constant B1	0.1	8.0
RT-model	Breakup time constant Cτ	1.5	0.95
	Size constant CRT	0.5	0.4

). Droplet evaporation followed the Frossling model, while droplet collisions were treated with the NTC model. Combustion was computed using the SAGE detailed chemistry solver, which was active from the start of injection until the end of combustion.

The KH-RT model describes droplet breakup by coupling Kelvin–Helmholtz (KH) instability with Rayleigh–Taylor (RT) instability, with the core governing equations as follows:

KH instability characteristic time:

$${\tau }_{KH}={\displaystyle \frac{\sqrt{{\rho }_l{d}_0}}{\sqrt{2\pi \sigma }}}\cdot {\displaystyle \frac{1}{W{e}^{0.5}}}$$

(1)

where ${\rho }_l$ is the droplet density, $d_0$ is the initial droplet diameter, is the surface tension, and $We$ is the Weber number.

RT instability characteristic time:

$${\tau }_{RT}=\sqrt{{\displaystyle \frac{d_0}{g({\rho }_l-{\rho }_g)/{\rho }_l}}}$$

(2)

where $g$ is the gravitational acceleration and ${\rho }_g$ is the gas-phase density.

The transport equations for turbulent kinetic energy (k) and its dissipation rate (ε) are as follows:

$${\displaystyle \frac{\partial (\rho k)}{\partial t}}+{\displaystyle \frac{\partial }{\partial x_j}}(\rho U_{j}k)={\displaystyle \frac{\partial }{\partial x_j}}[(\mu +{\displaystyle \frac{{\mu }_t}{{\sigma }_k}}){\displaystyle \frac{\partial k}{\partial x_j}}]+G_k-\rho \epsilon -Y_M+S_k$$

(3)

$${\displaystyle \frac{\partial (\rho \epsilon )}{\partial t}}+{\displaystyle \frac{\partial }{\partial x_j}}\left[\rho U_j\epsilon -[(\mu +{\displaystyle \frac{{\mu }_t}{{\sigma }_{\epsilon }}}){\displaystyle \frac{\partial \epsilon }{\partial x_j}}]\right]=\rho C_1{S}_{\epsilon }-C_2\rho {\displaystyle \frac{{\epsilon }^2}{k}}+C_{\epsilon 1}{G}_b+{\varphi }_{\epsilon }$$

(4)

where C₂ = 1.9 (dissipation term coefficient), C_ε1 = 1.9 (turbulence generation term coefficient), σ_k = 1.0 (Prandtl number for turbulent kinetic energy k), σ_ε = 1.2 (Prandtl number for dissipation rate ε), G_k is the turbulence kinetic energy generated by the velocity gradient of the laminar flow, S_k and Φ_ε are user-defined source terms, and Y_M represents turbulence fluctuations generated by compressibility through expansion diffusion, expressed as follows:

$$Y_M=2\rho \epsilon {M_t}^2$$

(5)

In which M_t denotes the turbulent Mach number.

During the numerical simulation process, the conservation laws of mass, momentum, and energy are strictly followed. The mass conservation equation is given as follows:

$${\displaystyle \frac{\partial }{\partial t}}\left({\rho }_m\right)+\nabla \cdot ({\rho }_m{U}_m)=0$$

(6)

where U_m is the mass-averaged velocity, and ρ_m is the mixture density.

The momentum equation is as follows:

$${\displaystyle \frac{\partial }{\partial t}}\left({\rho }_m{U}_m\right)+\nabla \cdot ({\rho }_m{U}_m{U}_m)=-\nabla p+\nabla \left[{\mu }_m\left(\nabla U_m+\nabla U_m^T\right)\right]+{\rho }_{m}g+F-\nabla \cdot \left({\displaystyle \sum {\partial }_k{\rho }_k{U}_{dr,k}{U}_{dr,k}}\right)$$

(7)

in which F denotes the body force, μm represents the mixture viscosity, and U_dr,k stands for the drift velocity of phase k.

The energy equation is given as follows:

$${\displaystyle \frac{\partial }{\partial t}}{\displaystyle \sum _k\left({\partial }_k{\rho }_k{E}_k\right)}+\nabla \cdot {\displaystyle \sum _k\left({\partial }_k{U}_k\left({\rho }_k{E}_k+p\right)\right)}=\nabla \cdot \left(k_{eff}\nabla T-{\displaystyle \sum _k{\displaystyle \sum _j{h}_{j,k}{J}_{j,k}}}+\left({\tau }_{eff}\cdot U\right)\right)+S_h$$

(8)

in which h_j,k denotes the enthalpy of species j in phase k, and J_j,k represents the diffusion flux of species j in phase k; S_h refers to the energy source term, k_eff stands for the effective thermal conductivity, and E_k is defined as follows:

$$E_k=h-{\displaystyle \frac{p}{\rho }}+{\displaystyle \frac{{u_i}^2}{2}}$$

(9)

For compressible flow, enthalpy h is a function of pressure and temperature, as follows:

$$dh={\left({\displaystyle \frac{\partial h}{\partial T}}\right)}_{p}dT+{\left({\displaystyle \frac{\partial h}{\partial p}}\right)}_{T}dp=c_{p}dT+{\left({\displaystyle \frac{\partial h}{\partial p}}\right)}_{T}dp$$

(10)

in which T denotes the temperature and c_p represents the specific heat at constant pressure. By adopting the volumetric thermal expansion coefficient (β), Equation (7) can be rewritten as follows:

$$\Delta h=c_p\Delta T+{\displaystyle \frac{1-\beta T}{\rho }}\Delta P$$

(11)

The parameter configurations of the KH-RT model are presented in Table 2, and the adoption of other submodels employed for the numerical simulation is compiled in

Table 3. Selection of simulation submodel.

Physical Model Name	Submodel Settings
Turbulence model	RNG k-ε [54]
Spray breakup model	KH-RT [55]
Drop turbulent dispersion model	Wall Film-O’Rourke [56]
Spray collision model	NTC collision [57]
Wall heat transfer model	Han and Reitz model [58]
Combustion model	SAGE [59]
Carbon smoke emission model	Hiroyasu soot [60]
NOx formation model	Extended Zeldovich [61]
Soot emissions	Hiroyasu soot model

.

2.4. Simulation Model Calibration

The employed grid refinement techniques allow independent control over duration and refinement levels and support parallel execution for precise grid sizing. Model validation was conducted against experimental data obtained from the baseline diesel operation of the prototype engine. Four base grid sizes (8 mm, 6 mm, 4 mm, and 2 mm) were evaluated. For the selected 4 mm base grid, a two-level Adaptive Mesh Refinement (AMR) was applied to resolve the boundary layer, while the fuel spray region was refined using a three-level fixed embedding strategy. The resulting mesh near the top dead center (TDC), detailing the spray and near-wall regions, is shown in Figure 4.

Experimental data for model validation were provided by Zhenjiang CSSC Marine Power Co., Ltd. (Zhenjiang, China). The simulation model was calibrated against the measured in-cylinder pressure trace from engine bench tests. As shown in Figure 5, under the reference operating condition (engine speed: 1500 rpm, injected fuel mass: 70 mg, IMEP: 12.63 bar), the simulated pressure curve demonstrates excellent agreement with the experimental data. The error is minimal, and key features such as the timing and magnitude of the pressure rise are accurately reproduced. Based on a prior grid independence study and considering computational efficiency, a base grid size of 4 mm was selected for all subsequent simulations in this work.

Initial conditions are set as follows: for the naturally aspirated operating condition, the initial in-cylinder pressure is 1 bar and the temperature is 350 K, referencing typical initial states of marine engines from idle to medium load; for the turbocharged operating condition, the initial pressure is 2 bar, with temperatures of either 350 K or 500 K when coupled with heating. Initial turbulence parameters are calibrated based on engine geometric parameters, including a bore diameter of 105 mm and a stroke of 125 mm, referencing the application of the RNG k-ε model in engines with similar bore diameters (Han & Reitz, 1995) [54]. Temperature boundary conditions are established with a cylinder wall temperature of 373 K, a piston crown temperature of 423 K, and a cylinder head temperature of 453 K, based on experimentally measured engine thermal boundary conditions. For the baseline operating condition, the fuel mass is determined by total energy input based on the diesel baseline (70 mg/cycle, LHV = 42.8 MJ/kg). During the validation phase, the peak in-cylinder pressure error between simulation and experiment is 3.8%, the IMEP error is 4.2%, and the peak heat release rate error is 5.1%, all within the engineering acceptable range of less than 6%.

3. Results and Discussion

3.1. Analysis of the Impact of AEPs on Combustion Characteristics

Figure 6 shows the in-cylinder pressure and heat release rate curves under various AEPs. AEP refers to the proportion of the total heat value of ammonia fuel to the total fuel energy input into the cylinder. The computational results indicate that as the AEP increases, the peak cylinder pressure continuously decreases. The ammonia start of injection (SOI) timing of −10 °CA was selected because this crank angle corresponds to the peak in-cylinder temperature after n-heptane combustion (see Figure 7, temperature curve for the AEP = 70% operating condition), enabling maximum utilization of the thermal environment to promote ammonia ignition. An additional comparative operating condition with SOI = −7 °CA was tested, and the results showed an 8.3% improvement in ammonia combustion efficiency at SOI = −10 °CA, leading to the adoption of this timing. When the AEP rises to 70%, misfire occurs in the cylinder, meaning that the fuel cannot be stably ignited and combusted.

From the heat release rate curves, it can be observed that in the thermal atmosphere compression ignition combustion mode, the overall combustion and heat release process in the cylinder is divided into two stages: the compression ignition and combustion stage of n-heptane, and the diffusion combustion stage of ammonia. The timing of n-heptane compression ignition is influenced by the thermodynamic conditions in the cylinder. Under constant initial pressure and temperature conditions, the timing of n-heptane compression ignition remains consistent across different AEPs. However, due to the varying masses of n-heptane at different AEPs (higher AEPs result in less n-heptane injected and less n-heptane entering the cylinder per cycle), the energy released during combustion differs, leading to different heating effects. This is visually evident in the differing magnitudes of increase in the average in-cylinder temperature after n-heptane combustion, as shown in Figure 7. Figure 7 also reveals that the average in-cylinder temperature at the ammonia injection timing (−10 °CA) decreases as the AEP increases.

When the AEP reaches 70%, misfire occurs with the ammonia fuel, indicating that the amount of n-heptane fuel is unable to generate enough heat to ignite the ammonia. The reason for choosing −10 °CA as the ammonia injection timing is that the in-cylinder temperature is close to its peak at this point, as can be seen from the in-cylinder temperature curve under the AEP = 70% (misfire) condition in Figure 7. Therefore, injecting ammonia at −10 °CA allows it to enter the cylinder at the highest temperature, fully utilizing the thermal atmosphere conditions generated by n-heptane combustion, thereby achieving as high an AEP as possible.

Figure 8 illustrates the consumption process of n-heptane, where the left axis represents the ratio of the remaining unburned n-heptane mass to the total n-heptane mass. Thus, the left axis, ranging from 1 to 0, indicates the transition from n-heptane not igniting to n-heptane being completely combusted. Through a corresponding analysis with the heat release rate curve, it can be observed that the combustion process of n-heptane can be divided into two stages: the first stage is characterized by the concentrated detonation of n-heptane, where it undergoes intense combustion after reaching the ignition conditions, resulting in a sharp decrease in n-heptane mass; the second stage is the slow combustion of n-heptane, where after the concentrated detonation, the remaining n-heptane burns at a significantly reduced rate, exhibiting a state of slow oxidation with a very low heat release rate.

The slowdown in n-heptane combustion during the second stage is related to the equivalence ratio. After the first stage of combustion, the equivalence ratio of n-heptane in the cylinder drops instantaneously. Since the chemical reaction rate is dependent on the reactant concentration, the decrease in the equivalence ratio leads to a significant reduction in the reaction rate. As the AEP increases, the total mass of n-heptane in the cylinder decreases, resulting in a longer duration for the second stage. This is because after the first stage of n-heptane combustion, the equivalence ratio in the cylinder is even lower, leading to a further decrease in combustion and reaction speeds.

Moreover, as the AEP increases, the initial combustion speed of ammonia fuel becomes slower due to the lower in-cylinder temperature. Therefore, after the AEP increases, the heat release rate does not rise rapidly from the moment of ignition; instead, it undergoes a period of slow heat release before significantly increasing.

Figure 9 depicts the in-cylinder temperature field distribution and flame development process under different AEPs. As shown in Figure 9, with an increase in the AEP, the reduction in n-heptane in the cylinder leads to a gradual decrease in the in-cylinder temperature before ammonia ignition. Consequently, when the AEP reaches 70%, misfire occurs in the cylinder.

Observing the flame development process at a 50% AEP, it can be seen that after ammonia is injected into the cylinder, it mixes with the air to form a combustible mixture and undergoes slow oxidation (low-temperature combustion), causing the in-cylinder temperature to further rise, corresponding to the stage of slow heat release rate increase. When the in-cylinder temperature reaches a certain level, the ammonia fuel in the cylinder undergoes concentrated detonation, and the ammonia spray ignites and burns instantaneously, leading to a sharp increase in the heat release rate. A high-temperature combustion zone forms along the ammonia spray distribution area (as shown at 0 °CA for the 50% AEP).

The compression ignition combustion process of ammonia under hot conditions shares similarities with the combustion characteristics of traditional diesel engines. However, ammonia compression ignition combustion has a longer ignition delay period, resulting in a different heat release pattern compared to diesel engines. As the AEP increases, the further extension of the ammonia ignition delay period leads to significant changes in the fuel distribution within the cylinder during ammonia combustion. Therefore, the combustion characteristics and flame development of ammonia undergo changes.

3.2. Effects of AEP on Combustion Characteristics at Various Compression Ratios (CRs)

From the analysis results presented above, it is evident that the primary reason for ammonia’s failure to ignite and combust in the cylinder as the AEP increases is the reduction in n-heptane, which leads to inadequate heat provision from n-heptane and subsequently low in-cylinder temperatures, preventing ammonia from igniting spontaneously. When the piston moves upwards towards the TDC, the in-cylinder temperature is primarily determined by two factors: the amount of heat released from n-heptane combustion and the influence of the compression ratio on the pressure after compression at TDC. A higher compression ratio results in higher compression pressure and temperature at TDC after compression. Therefore, increasing the compression ratio is one of the effective measures to address the difficulty of igniting low-reactivity fuels through compression. This section focuses on studying the impact of the compression ratio on ammonia’s spontaneous ignition and combustion characteristics.

The original engine has a CR of 16. Based on this, the compression ratio is increased to 18 and 20, respectively. The influence of increasing AEPs on ammonia combustion under different CRs is calculated and compared with that under the original compression ratio. This analysis aims to investigate the effect of increasing the compression ratio on combustion characteristics at various AEPs.

Figure 10, Figure 11 and Figure 12 compare the variations in in-cylinder pressure, heat release rate, and n-heptane mass fractions under different CRs. The results indicate that the CR has a significant impact on the spontaneous ignition of n-heptane, compression ignition of ammonia, and diffusion combustion characteristics under hot conditions.

When the CR is increased from 16 to 18, there is a noticeable change in the heat release pattern, with more concentrated heat release and an increase in the maximum in-cylinder pressure. At a CR of 18, a more pronounced heat release is observed at a 70% AEP. Further increasing the CR to 20 results in stable ignition and efficient combustion at the 70% substitution rate. Therefore, increasing the CR is an effective way to enhance the AEP. At a higher CR, the fuel burns more rapidly and the heat release is more concentrated, which favors efficient combustion of ammonia fuel.

However, it is important to note that as the CR increases, both n-heptane and ammonia burn more rapidly. Excessively fast combustion can adversely affect engine reliability. In this combustion mode, n-heptane is ignited through homogeneous charge compression ignition (HCCI), which involves concentrated ignition and rapid detonation during the piston’s upward stroke. As a result, the rapid detonation and heat release of n-heptane at higher CRs can easily lead to engine knocking, increasing the mechanical load on the engine and potentially causing damage to mechanical components or reducing air tightness, thereby shortening the engine’s lifespan and reducing its reliability.

The calculation results show that when the engine’s CR is increased to 20, the maximum in-cylinder pressure rises to about 14 MPa under moderate load conditions. If the engine’s operating load is further increased, the peak pressure will be even higher, approaching the engine’s pressure limit. Therefore, there are certain limitations to increasing the CR.

Figure 13 analyzes the impact of increasing the CR on the ignition and combustion heat release characteristics of n-heptane and ammonia. As shown in the figure, the effect of the CR on the ignition and combustion of n-heptane is mainly manifested in the significant advancement of the n-heptane ignition timing due to the rapid increase in in-cylinder compression pressure and temperature as the CR rises. However, there is no noticeable difference in the first-stage combustion heat release rate of n-heptane.

The increase in the CR also has a significant impact on the heat release pattern of ammonia. It can be observed that as the CR increases, the ignition speed of ammonia becomes faster, and the heat release becomes more concentrated.

Overall, the increase in the CR has a clear influence on the ignition characteristics, specific combustion heat release patterns, and combustion characteristics of both n-heptane and ammonia. Increasing the CR is beneficial for stable ignition and efficient combustion of ammonia.

Figure 14 and Figure 15 illustrate the influence of varying CRs on the in-cylinder temperature field distribution under conditions of 50% and 60% AEP, respectively. The in-cylinder temperature field distribution reflects the combustion process and characteristics of the fuel within the cylinder. It can be observed that as the CR increases, the temperature distribution within the cylinder at the moment of ammonia fuel injection rises significantly, leading to faster ignition of ammonia combustion.

However, following the ignition of ammonia fuel, there are minimal differences in the temperature distribution within the cylinder across different CRs. This is because the diffusion combustion of ammonia is primarily governed by the injection process of the ammonia fuel. Although the timing of ammonia ignition and combustion varies with different CRs, the overall characteristics of ammonia diffusion combustion remain consistent due to the unchanged injection timing and strategy of ammonia, resulting in no significant differences in the temperature distribution characteristics within the cylinder. Thermal NO_x (accounting for approximately 40–55%) is calculated using the Extended Zeldovich model, while fuel NO_x (accounting for about 45–60%) is generated from the nitrogen atoms in ammonia. As the ammonia energy percentage (AEP) increases from 50% to 80%, the proportion of fuel NO_x rises from 45% to 60% due to the increased ammonia input (see Figure 16 and Figure 17).

Figure 16 and Figure 17 present the characteristics and distribution patterns of in-cylinder NO_x generation under different CRs. The results indicate that an increase in the CR has a significant impact on NO_x generation and its distribution within the cylinder. This is because NO_x formation is strongly correlated with temperature; NO_x is primarily generated under high-temperature and oxygen-rich conditions. Therefore, as the cylinder temperature rises, NO_x formation becomes more favorable, and consequently, an increase in the CR affects NO_x production.

Furthermore, the results reveal notable differences in NO_x generation and distribution characteristics within the cylinder under varying AEPs. During ammonia combustion, apart from the thermal NO_x formed through the oxidation of nitrogen at high temperatures, there is also fuel-NO_x generated from the nitrogen contained within the ammonia fuel. As the AEP increases, the amount of ammonia fuel injected into the cylinder rises, leading to an increase in fuel-NO_x generation. This, in turn, alters the NO_x generation and distribution characteristics within the cylinder.

3.3. Analysis of the Impact of Intake Boundary Conditions Under Various AEPs

3.3.1. Analysis of the Impact of Intake Boosting on Combustion Characteristics

In addition to increasing the CR, increasing intake pressure and intake temperature are also effective measures to promote the ignition and combustion of fuel within the cylinder. Previous analysis focused on naturally aspirated conditions; here, we investigate the influence of intake boundary conditions on ammonia combustion characteristics under varying AEPs by employing intake boosting and intake heating separately. Furthermore, we explore control strategies for intake boundary conditions that can effectively enhance AEPs by coupling intake boosting and intake heating. Although the relative sensitivity coefficients of the heat release rate (HRR) to variations in the equivalence ratio are similar under 1 bar and 2 bar intake pressures, the absolute magnitude of HRR changes is greater at higher intake pressures. Additionally, due to the higher in-cylinder air density under elevated pressure, even minor fluctuations in the equivalence ratio have a more pronounced impact on fuel mixing uniformity, resulting in stronger sensitivity of actual combustion stability to variations in the equivalence ratio.

An initial comparative analysis was conducted on the cylinder pressure and heat release rate calculations under initial pressures of 1 bar and 2 bar, as illustrated in Figure 18. The results reveal that the in-cylinder pressure and combustion heat release rate are influenced by the initial in-cylinder pressure, leading to distinct combustion characteristics.

Firstly, an increase in the initial in-cylinder pressure has a significant impact on the ignition and combustion of n-heptane. Under an initial pressure of 2 bar, the combustion rate of n-heptane is markedly higher than that under 1 bar, resulting in a substantial increase in the peak heat release of n-heptane combustion. In addition to a notable increase in the peak heat release during the first stage of n-heptane combustion, the second stage of combustion also exhibits a significant acceleration and advancement. Furthermore, it is observed that under the initial pressure of 2 bar, the heat release rate of n-heptane combustion is more sensitive to the equivalence ratio (the concentration of the n-heptane premixture). As the AEP increases, the proportion of n-heptane in the intake decreases, leading to a lower equivalence ratio of n-heptane, and consequently, a reduced heat release rate of n-heptane combustion, as depicted in Figure 18.

Secondly, the initial pressure also exerts a notable influence on the ammonia diffusion combustion process. An increase in the initial pressure implies an increase in the intake pressure, which in turn enhances the in-cylinder air intake per cycle, resulting in an increased air density within the cylinder. Under these conditions, the heating effect produced by the combustion of the same mass of n-heptane is diminished, leading to a decrease in in-cylinder temperature and subsequently making it more difficult for ammonia to ignite and combust. The computational results indicate that under an initial pressure of 2 bar, misfire occurs at an AEP of 60%, whereas at 1 bar, misfire is observed only when the AEP reaches 70%. This finding suggests that for the ammonia thermally assisted compression ignition (TACI) combustion mode, an increase in intake pressure is detrimental to the enhancement of the AEP. As the intake pressure increases, the density of the in-cylinder working medium rises, necessitating more heat release from n-heptane combustion to achieve the thermal conditions required for stable ammonia ignition and combustion.

Figure 19 analyzes the impact of different initial pressures on n-heptane combustion. As illustrated in Figure 19, although an increase in initial pressure makes it more challenging to achieve a higher AEP, it favorably influences the ignition and complete combustion of n-heptane, thereby reducing incomplete combustion losses and enhancing combustion efficiency. The computational results reveal a significant improvement in n-heptane combustion efficiency when the initial pressure is increased from 1 bar to 2 bar.

The combustion efficiency of n-heptane is influenced by the n-heptane premixed equivalence ratio. Under low intake pressure conditions, as the AEP increases (leading to a decrease in the n-heptane premixed equivalence ratio), the combustion efficiency of n-heptane decreases notably. Specifically, at a 70% substitution rate, the combustion efficiency of n-heptane is approximately 92%; at an 80% substitution rate, it drops to 80%; and at a 90% AEP, only about 67% of the n-heptane is combusted. However, when the initial pressure is increased to 2 bar, the combustion efficiency of n-heptane at a 90% AEP rises to approximately 93%, and the combustion efficiencies for lower AEPs approach 100%. This demonstrates that increasing the intake pressure has a profound effect on enhancing combustion efficiency.

Figure 20 compares the variations in the n-heptane mass fraction and the heat release rate curves under different initial pressures. Through comparison, it is evident that the initial pressure has a significant influence on the ignition timing of n-heptane, combustion speed, and the diffusion combustion process of ammonia. Generally, increasing the intake pressure enhances the combustion speed of n-heptane, leading to earlier ignition. However, due to the increased intake density resulting from higher intake pressure, the intensity of the in-cylinder thermal atmosphere generated by n-heptane combustion is reduced, which slows down the ignition speed of ammonia, delays the combustion phase, and reduces the heat release rate.

As illustrated in Figure 21 and Figure 22, which show the temperature field distribution and NO generation distribution under different initial pressures, it can be observed that an increase in intake pressure results in a notable decrease in the combustion temperature within the cylinder, along with a significant reduction in NO generation. Additionally, there are substantial changes in the distribution characteristics of both temperature and NO within the cylinder. Figure 22 illustrates the distribution characteristics of NO (the primary component of NO_x, accounting for over 80%), while Figure 16 and Figure 17 depict total NO_x emissions (including NO, NO₂, and fuel NO_x). These figures, respectively, correspond to the distribution analyses of the “key component” and “total emissions,” providing complementary insights into the effects of compression ratio and intake pressure on NO_x formation.

3.3.2. Analysis of the Impact of Intake Air Heating on Combustion Characteristics

Based on the analysis above, it is evident that increasing the intake pressure is not conducive to enhancing the AEP. This is because an increase in intake pressure leads to a higher density of the working medium within the cylinder, making it more difficult to achieve the thermal conditions necessary for ammonia ignition and combustion. Consequently, this section explores methods to improve the AEP by increasing the intake temperature.

Figure 23 illustrates the cylinder pressure and heat release rate at different intake temperatures. According to the computational results, increasing the intake temperature does not resolve the misfire issue at a 70% AEP. This is primarily because, as the intake temperature rises, the amount of fresh air entering the cylinder decreases, leading to an insufficiency of oxygen required for fuel combustion within the cylinder. The heat release rate curve indicates that, with an increase in intake temperature, the heat release rate of n-heptane combustion decreases, suggesting that n-heptane combustion is severely affected.

This conclusion is also supported by Figure 24, which shows that n-heptane becomes more difficult to completely combust as the intake temperature increases. At a 70% AEP, the combustion efficiency of n-heptane significantly decreases under initial temperature conditions of 400 K. This indicates that, under conditions of oxygen deficiency, n-heptane combustion is more susceptible to the impact of ammonia injection (due to the high latent heat of vaporization of ammonia, which absorbs a substantial amount of heat when injected into the cylinder, thereby lowering the cylinder temperature). As the AEP further increases, n-heptane combustion rapidly deteriorates. At a 90% AEP, the combustion efficiency of n-heptane is approximately 67% at an initial temperature of 350 K, but it drops to around 28% at an initial temperature of 400 K.

Figure 25 compares the variations in the n-heptane mass fraction and the combustion heat release rate curves under different intake temperatures for AEPs of 50% and 60%. Through comparison, it can be observed that as the intake temperature increases, although a higher compression temperature provides better ignition conditions (earlier n-heptane ignition), the combustion speed of n-heptane decreases due to the reduced oxygen content. Consequently, as the cylinder continues to dissipate heat, the heating effect of n-heptane combustion on the cylinder interior diminishes, leading to a delay in the ignition timing of ammonia. However, the combustion heat release rate of ammonia fuel is not significantly affected.

Figure 26 and Figure 27 compare the computational results of in-cylinder temperature distribution and NO distribution under different AEPs and intake temperatures. The results indicate that intake temperature has a certain impact on both the in-cylinder temperature distribution and NO distribution. As observed in Figure 26, when the initial temperature is increased from 350 K to 400 K, the in-cylinder temperature at the moment of ammonia injection actually decreases. This finding supports the previous analysis that the deteriorated combustion of n-heptane leads to a reduced heating effect. Furthermore, with changes in the thermodynamic conditions within the cylinder and the combustion heat release characteristics of ammonia fuel, there are noticeable variations in the generation and distribution patterns of NO within the cylinder.

3.3.3. Synergistic Effect of Intake Turbocharging and Intake Heating on Enhancing AEPs

The detrimental effect of increased intake temperature on the combustion of n-heptane and ammonia fuels primarily stems from the reduction in oxygen content associated with this temperature rise. Therefore, to ensure sufficient intake air quantity when elevating the intake temperature, it is necessary to correspondingly increase the intake pressure, thereby providing the necessary air for the combustion of n-heptane and ammonia. Consequently, this section discusses the impact of coupling intake pressure boosting and intake air heating on ammonia-fueled compression ignition under hot conditions.

Figure 28 presents the computational results of combustion characteristics under various AEPs at an initial pressure of 2 bar and an initial temperature of 500 K. As evident from Figure 28, the coupling of intake pressure boosting and intake air heating significantly promotes the ignition and combustion of ammonia fuel. Under these conditions (2 bar intake pressure and 500 K intake temperature), the AEP can be increased to 80%, achieving stable ignition and efficient combustion. Across different AEPs, the combustion efficiency of n-heptane remains close to 100%, although the specific heat release profiles vary depending on the AEP.

Under initial conditions of 2 bar and 500 K, an analysis was conducted on the in-cylinder temperature distribution and NO generation distribution across different AEPs. The computational results presented in Figure 29 and Figure 30 indicate that there are distinct differences in the in-cylinder temperature distribution under varying AEPs. These differences primarily arise from the different heat release patterns of ammonia combustion. As observed in Figure 28, under different AEPs, there are significant variations in the combustion heat release rates and heat release profiles of both n-heptane and ammonia, leading to differences in the in-cylinder temperature distribution. When the compression ratio (CR) increases from 16 to 20, the total NO_x emissions rise from 8.2 g/kWh to 11.5 g/kWh (a 40.2% increase) under the AEP = 50% operating condition, and from 9.8 g/kWh to 13.3 g/kWh (a 35.7% increase) under the AEP = 70% operating condition. This indicates that the higher compression ratio significantly increases total NO_x production by elevating in-cylinder temperatures, while the increased proportion of fuel NO_x at higher AEP levels mitigates the rate of increase. The formation of N₂O primarily stems from the incomplete oxidation of ammonia (pathway: NH₃ → NH₂ → N₂O). Simulation results indicate that N₂O emissions are 0.8 g/kWh at an ammonia energy percentage (AEP) of 50%, increasing to 1.5 g/kWh at an AEP of 80% due to the higher ammonia input. Coupling intake air heating (500 K) can reduce N₂O emissions by approximately 22%, as elevated temperatures promote the decomposition of N₂O. Furthermore, due to the close correlation between NO_x generation and temperature, there are also notable differences in NO generation. In summary, increasing the CR or simultaneously increasing the intake pressure and intake temperature are effective means of enhancing the AEP.

4. Conclusions

This study establishes a dual-fuel combustion simulation model for ammonia/n-heptane, integrating a chemical reaction kinetics model to delve into the combustion characteristics of ammonia under thermally active conditions. An in-depth analysis is conducted on the combustion behaviors across various AEPs, along with an investigation into the influence of the compression ratio, intake boundary conditions, and ammonia injection strategies on in-cylinder combustion characteristics and NO_x emissions. Employing direct ammonia injection within the cylinder, which boasts higher combustion and thermal efficiencies, this research provides foundational insights and guidance for future studies on ammonia fuel combustion, holding significant theoretical value and practical application potential. The key findings of this study are summarized as follows:

(1)

Under the thermally active conditions generated by n-heptane compression ignition, ammonia fuel can achieve stable ignition and efficient combustion. The combustion characteristics of direct ammonia injection compression ignition closely resemble those of conventional diesel combustion. The direct-injection method facilitates ammonia diffusion combustion, yielding high combustion and thermal efficiencies while minimizing unburned ammonia emissions.

(2)

An analysis of the impact of AEPs on combustion characteristics reveals that as the AEP increases, the ammonia ignition and combustion characteristics are markedly affected. At an AEP of 70%, significant misfire occurs, attributed to ammonia’s strong resistance to auto-ignition and its high evaporation latent heat.

(3)

Elevating the compression ratio has a profound effect on ammonia ignition and combustion performance, with notable changes in ammonia’s combustion characteristics and heat release patterns observed across different CRs. Increasing the compression ratio effectively raises the proportion of ammonia fuel; at a compression ratio of 20, a 70% ammonia fuel ratio can achieve stable ignition and efficient combustion.

(4)

Increasing intake pressure or temperature alone does not effectively enhance the AEP. This is because elevated intake pressure increases in-cylinder working fluid density, diminishing the thermal atmosphere generated by n-heptane combustion, while increased intake temperature reduces the intake air volume, impacting n-heptane ignition and combustion efficiency and similarly weakening the in-cylinder thermal atmosphere. The results indicate that combining intake boosting and heating can effectively increase the AEP, further reducing carbon emissions from ammonia-fueled engines.

(5)

This strategy (port injection of n-heptane + direct in-cylinder injection of ammonia) enables stable combustion at medium-to-high average effective pressure (AEP). However, it faces issues such as ignition delay fluctuations, high sensitivity of high AEPs to injection parameters, and the core contradiction between ammonia slip and ignition. Compared with traditional direct diesel injection, this strategy is more suitable for medium-to-high load low-carbon marine propulsion. Optimizing the compression ratio (CR = 20) and intake coupling (2 bar + 500 K) can help manage these contradictions. In the future, it can be further optimized through the use of variable nozzles, pre-chambers, or hydrogen blending (5–10%).