Study of Effects on Performances and Emissions of a Large Marine Diesel Engine Partially Fuelled with Biodiesel B20 and Methanol

Abstract

The impact of fossil fuel utilisation in different combustion systems on climate change due to greenhouse gas accumulation in the atmosphere is rather evident. A part of these gases comes from the large engines used for propulsion in marine applications. In the continuous global effort made by engine manufacturers to mitigate this negative impact, one way is represented by the utilisation of alternative fuels such as biodiesel and methanol, based on dedicated research to fulfil the more stringent regulations concerning pollutant emissions issued by piston heat engines. In this study, a numerical investigation was conducted on a four-stroke large marine diesel engine (ALCO 16V 251C) at several engine speeds and full load conditions. Different blends of diesel–methanol and biodiesel B20–methanol with methanol mass fractions of 10% and 20% were considered for theoretical analysis in two techniques of methanol supply: direct injection mode of a blend of base fuel diesel/biodiesel B20 with methanol and injection of methanol after the intercooler, and direct injection of the base fuel. The results show that, if 10% in power loss can be acceptable, then for diesel–methanol 10%, in the direct injection technology, the NOx emission can be reduced up to 19%, but with a compromise of an 8% increase in SOOT emission, while for biodiesel B20–methanol 10%, with the same direct injection method, the NOx emissions increase by up to 58% with the benefit of reducing SOOT by up to 23% relative to the original diesel fuel operation. For a 20% methanol fraction in blend fuel, the drop in power is more than 10% regardless of the method of methanol supply and the base fuel, diesel, or B20 used.

1. Introduction

The increasingly intense climate changes in recent decades, amplified by the greenhouse effect of combustion gases from various fields of activity, have made humanity aware of some consequences of human activities on the environment. In response to these effects, researchers are looking for the identification of new fuels that are more environmentally friendly, while the legislation provides increasingly strict regulations on the emissions resulting from fossil fuel combustion. Among other alternative fuels, methanol and biodiesel can be considered promising future fuels for the defossilisation of internal combustion engines (ICEs) [1]. Until the large-scale implementation of new propulsion technologies, compression ignition engines that contribute to the prosperity of the world economy remain indispensable in the short and medium term due to their high reliability, efficiency, and flexibility, even if the emissions of these engines are considered one of the most important causes of the greenhouse effect, namely air contamination. It has been proven that these pollutants affect ecological systems and some components of these emissions could cause diseases such as lung diseases, asthma, and even cancer, which place the health of human beings at risk [2,3,4]. The frequent variations in the price of crude oil on the international market and the reduction in these reserves harm the predictability of the economy of most countries and impose the need to find stable and more environmentally friendly alternatives.

The utilisation of alternative fuels like hydrogen, natural gas, biodiesel, and alcohols can reduce both the consumption of petroleum-derived fossil fuels and their emissions. Primary alcohol fuels are promising fuels for diesel engines among other fuels like HVO, biodiesel, and e-fuels, due to their fundamental properties, a reduced stoichiometric air/fuel ratio, a raised oxygen content, a higher H/C ratio, a lack of C–C bonds, the presence of oxygen in their molecules [5], and a lower viscosity. Methanol is not usually used in spark ignition (SI) engines and compression ignition (CI) engines but it is researched to be a substitute for conventional diesel fuel due to its potential to improve the combustion process and reduce CO, HC, and PM emissions [6,7]. In the diesel/methanol dual-fuel mode, methanol is utilised as the low-reactivity fuel, while diesel is utilised as the high-reactivity fuel [8]. Due to the fact that methanol contains oxygen atoms (five-fold higher), the oxygen available for combustion is increased. In addition, methanol’s heat of vaporisation is higher compared to that of diesel fuel; thus, the use of methanol in CI engines can reduce both NOx and smoke [9,10].

Utilising primary alcohol fuels brings many advantages by replacing conventional fuels, like being colourless, a pure substance, obtainable from renewable and conventional sources like fossil fuel, including biomass, coal, and natural gas [6]. Another major advantage is that primary alcohols can be stored in a liquid state at standard ambient temperature and pressure; thus, storage and distribution can be handled with ease [1]. Due to the low cetane numbers, low viscosity, and low lubricity, it is very difficult to use primary alcohols in CI engines without ignition assistance. In general, the volatility, low surface tension, and viscosity strongly affect the distribution of the fuel droplet size, the atomisation quality of the fuel injection, and the uniformity of the mixture [6]. The relatively high vapor pressure and low viscosity of pure methanol causes it to volatilise readily into the air, which leads to the cold-start problem [6,11,12].

Table 1. Physico-chemical properties of different fuels [13,14,15,16,17,18,19,20,21,22,23,24,25,26].

Properties	Diesel Fuel	Biofuel	Biodiesel B20	Methanol
Chemical formula	C12H26-C14H30	C19H36O2	C15H23O	CH3OH
Molecular weight (g/mol)	170–220	292.6	219	32
Density @ 20 °C (kg/m3)	810–880	887	857.4	790
Boiling Point (°C)	125 i–400 f	330	200	64.7
Viscosity (20 °C) (cSt)	3.35	8.06	5.12	0.59
Flash Point (°C)	65–88	140	85	11
Autoignition temperature (°C)	204–340	380	-	470
Cetane number	40–55	55–56	52.5	3–5
Air/Fuel ratio at stoichiometric	14.7	12.6	14.2	6.5
Lower heating value (MJ/kg)	42.5	38.8	40.5	19.7
Heat of vaporisation (kJ/kg)	260	350	277	1175
Carbon content (%wt)	87	76.9	82.2	37.5
Hydrogen content (%wt)	12.6	12.4	10.5	12.5
Oxygen content (%wt)	0.004	10.7	7.3	50
Water content (mg/kg)	50	300	120
Carbon residue in %	0.001	0.1	0.032
Ash (% by mass)	0.016	0.087	0.023
Flame temperature (°C)	2054	-	-	1890

ⁱ—initial boiling point; ^f—Final boiling point.

presents a diversity of fuel properties including methanol.

The method used to supply methanol to the engine has a strong impact on the methanol fraction that can be used in the mixture with conventional diesel fuel. Methanol may be supplied in the CI engine using the following techniques [27]: diesel/methanol compound combustion (DMCC) is one of the most popular methods wherein methanol is injected at low pressure into the intake manifold and conventional diesel fuel is directly injected into the combustion chamber [8,9,28]. This method is also called the fumigation approach. A dedicated system is required for the injection of methanol. A vaporiser is desirable for fumigation, which involves further modification of CI engines. A maximum proportion of 50% of methanol can replace conventional diesel using this method [27]. Dual injection: In this method, a dedicated fuel supply and injection system are essential for injecting methanol (lower reactivity) during the intake stroke. By using this method, the replacement of diesel with methanol is up to 90%. Mixtures of diesel–methanol: Instead of diesel, a blend of diesel–methanol is directly injected into the engine cylinders. Compression ignition engines do not have any special requirements or hardware modifications to function with methanol–diesel blends [29]. Alcohol–diesel emulsion: To prevent phase separation between the two considered fuels, an emulsifier is prepared. Diesel can be replaced in proportion up to 25% using this technology [27]. The results indicate that the brake-specific fuel consumption (BSFC) diminished, NOx and smoke emissions were reduced, and CO and HC emissions increased in the diesel/methanol dual-fuel mode.

With the increase in methanol quantity, the ignition delay was increased but the combustion duration was reduced, accelerating the diffusion combustion [30]. At high loads, the increased noise of the combustion process provided the limit of the operating range. It has been found that a higher rate of exhaust gas recirculation can limit the speed of the heat release rate [12]. The dual-fuel RCCI (reactivity controlled compression ignition) combustion technology is the first step in the combustion process of the CI engine and the homogeneous charge compression ignition (HCCI) engine. The RCCI combustion method has as its disadvantage the high energy substitution ratio (ESR) which deteriorates the combustion process at low loads due to the misfire occurrence increasing the amount of unburned fuel in the exhaust gas. A second disadvantage is identified at high loads due to the increased level of noise that settles the upper limit of the engine operation. Another disadvantage is represented by an increase in HC and CO emissions due to the trapped unburned fuel in the crevice and squish region of the cylinder [31].

Biodiesel is one of the renewable fuels that has captured the attention of researchers. This fuel is considered to be utilised as a suitable and sustainable alternative to the conventional diesel fuel. Biodiesel has a slightly reduced lower heating value when compared to conventional diesel fuel, but its benefits consist in the reduction in pollutant emissions [32]. Compared to diesel, biodiesel has the advantages of a higher cetane number, absence of sulphur and aromatic compounds, excellent lubricity, low volatility, and an oxygen content of 11%, improving the oxidation of the SOOT and hydrocarbon emissions [33]. Numerous studies indicate a reduction in NOx and PM emissions with the use of alcohol and biodiesel blends at the cost of increasing CO emissions when compared to diesel and biodiesel. Researchers have obtained mixed results in the case of HC emissions. Research shows that HC emissions for alcohols are higher for biodiesel and lower for diesel [34,35], but other studies emphasise the importance of the proportion of alcohol and its effects over the CO and HC emissions. In the particular case of blends with methanol proportions of 10% and 15%, HC and CO emissions increase when compared to diesel and biodiesel [34]. Biodiesel–alcohol blends have a negative impact on BSFC. Exhaust temperature does not show major changes, presenting a similar behaviour for all of the fuels.

The use of biodiesel–methanol is going to cause a delayed start of combustion and a shorter combustion duration when compared with biodiesel. The maximum reduction in combustion duration is 12 deg. CA at 80% engine load when the engine is fuelled with biodiesel–methanol mixtures [36]. Chen et al. [30] investigated the effect of energy substitution ratio on the combustion characteristics of a dual-fuel engine fuelled with diesel/methanol, diesel/ethanol, and diesel/n-butanol [37]. The authors found that, with the increase in the methanol proportion, the heat release rate will increase and the combustion duration will diminish. Due to the stability of the mixture of methanol and diesel, the methanol replacement percentage cannot reach a high level [37]. In an attempt to mitigate the obstacles that occurred while blending diesel and methanol, many researchers have considered for their investigation the dual-fuel method with the injection of methanol in the intake manifold. By injecting methanol in the intake manifold and due to its higher latent heat of vaporisation, the temperature in the intake will decrease, increasing the fresh air charge density.

The advantage of the oxygen content of methanol is to improve the combustion process, resulting in an increase in the thermal efficiency by 1.9% [37]. Cheng et al. [38] compared the performance and emissions of a CI engine fuelled with biodiesel and methanol in blended and fumigation modes. They experimentally observed that, with a 10% methanol fraction, carbon dioxide (CO₂) emissions reduced by 2.5% and the NOx indicated a more than 5% reduction compared with diesel fuel. Therefore, the fumigation mode will cause a substantial increase in NO₂ pollutants in the exhaust gas, which is a disadvantage in comparison with the other methods of fuelling the engine. It was also found that the blended mode tended to improve the brake thermal efficiency of the engine at low engine loads, whereas, at high engine load conditions, the cooling effect of methanol reduced the efficiency. However, an opposite effect was observed utilising the fumigation method, reducing the efficiency at low engine load conditions but improving the efficiency at medium and high engine loads due to the better combustion process of methanol [39].

Qi et al. [40] analysed the combustion process, pollutant emissions, and performance of a CI engine fuelled with a blend of diesel–biodiesel with 50% biodiesel and 50% diesel fuel (BD50) considered as baseline fuel. In addition to the base fuel, methanol was added as an additive in fractions of 5% (MBD5) and 10% (MBD10). Authors have found that both MBD5 and MBD10 indicated a substantial reduction in smoke emissions with a very small reduction in CO emissions. The results indicate a similar trend of HC and NOx emissions at full load conditions for MBD5 and MBD10 compared to that of BD50. At lower load conditions, the combustion was delayed in the case of both fuels in comparison with BD50.

Changchun Xu et al. [41] studied the effect of injecting methanol in three locations, pre-intercooler, post-intercooler, and intake manifold under different load and engine speeds. At engine medium and low load conditions, regardless of the injection position, with the increase in methanol fractions, the maximum firing pressure inside the cylinder decreased, the peak heat release rate moved backwards, and the NOx emissions were reduced. Under the same conditions, the ignition delay increased, the total energy consumption increased, and the HC and CO emissions increased. For an engine operating at high load conditions, while increasing the methanol, maximum firing pressure increased at first and then decreased. The total energy consumption has a similar behaviour, and the HC and CO emissions slightly increased while the NOx emissions decreased. Under medium and low load conditions, changing the location of the methanol injector could diminish the adverse effects of using methanol blends. In the situation of injecting methanol into the pre-intercooler, which utilised the high temperature of compressed air, the methanol vaporised effectively, reduced both the ignition delay and combustion duration, and reduced HC and CO emissions. When injecting methanol at the post-intercooler, HC emissions slightly increased under low load conditions, ignition delay was reduced under medium load conditions, and cylinder pressure was the same as in the case of diesel fuel. For engines running at high-speed and high-load conditions, HC and CO emissions were reduced, and NOx emissions increased mainly due to the high in-cylinder temperature. With the methanol injector installed at the intake manifold, the lowest in-cylinder pressure and slowest heat release rate were found at low-load conditions, with the maximum reduction in NOx emissions at high-speed and high-load conditions.

Yilmaz [35] conducted a comparative study on a single-cylinder diesel engine fuelled with tertiary fuel blends such as biodiesel–ethanol–diesel (BED) and biodiesel–methanol–diesel (BMD) blends. The author found that, with the addition of ethanol, the NOx emissions were increased, but the methanol addition led to a reduction in HC and CO emissions. The author has found that an optimum alcohol blend ratio is essential for achieve pollutant emission control, due to the fact that the blend ratio has a direct influence on the oxygen content of the blended fuel [42]. Zaglinskis et al. [43] investigated the effect on the performance of a CI engine fuelled with blends of diesel–biodiesel (BD) and mixtures of diesel–biodiesel–methanol with a methanol fraction of 10% as an additive (MBD). They revealed that, for BD blends, the BSFC was higher by 3.5% while it variated from 2 to 13% for MBD blends in comparison with diesel. The results indicated a positive effect on the emission pollutants with a reduction in CO by 8%, HC by 32%, and SOOT emissions by 22% for the BD. For the MBD blend, the reduction was of about 13% for CM, 18% for HC, and 45% for SOOT emissions. [42].

In [44], the use of 10% n-butanol blended with diesel was investigated on the NEDC cycle. The results indicated a reduction in PM and smoke emissions, while NOx and CO₂ did not register any changes with an increase in CO and HC emissions. The purpose of the study was to identify the trends in methanol utilisation when blended with conventional diesel fuel and biodiesel B20, as base fuel, when fuelling a large heavy-duty diesel engine dedicated to naval applications. For this study, a theoretical investigation was conducted, calibrating an engine model developed in the software tool AVL Boost 2022R1 of the ALCO 251C for naval applications using experimental data. The simulations were conducted by considering two methods of supplying methanol, direct injection mixed with base fuel, and indirect injection (fumigation) with direct injection of the base fuel.

The novelty of this study consists in an assessment of the effects on performances and emissions of a large marine diesel engine partially fuelled with biodiesel B20 and methanol, which was conducted by numerical simulation using a model developed in dedicated software, which was calibrated on experimental test bench data. It emphasizes that methanol fuel by direct injection or by fumigation modifies the engine performance and emissions. A comparative analysis was carried out between two different base fuels, diesel and biodiesel B20, and two methods of methanol fuelling.

2. Materials and Methods

2.1. Research Methodology

The steps of the research methodology for this study are presented in the roadmap below (Figure 1), and the descriptions of the steps are inserted in the text boxes.

2.2. Simulation Setup

The technical specifications of the diesel engine ALCO 251 C considered for these investigations are presented in

Table 2. Engine specifications.

Application	Marine
Brake Power	2243 kW
BMEP	14.8 bar
Rated speed	1050 rpm
BSFC@1050 rpm	220.32 g/kWh
Compression ratio	12.5
Stroke	266.7 mm
Bore	228.6 mm
Configuration	V16
Displacement	175.3 L
Turbocharger	ALCO 720
Turbocharger speed	18,600 rpm
Turbocharger pressure ratio	3
Fuel system	Mechanical, pump-injector system
Fuel pump timing	27.5°
Injector	9 holes × 0.35 mm
Injection pressure	260 bar
Dry weight	18,609 kg
Wet weight	20,175 kg
Valve timing	Fixed, Camshaft driven
Camshaft overlap	123°
Length	5029 mm
Width	1551 mm
Height	2438 mm

.

As agreed upon, comparative tests were carried out with the HS 5800 turbo from Hispano Suiza, currently Hedemora Turbo & Diesel, to the original ALCO 720 turbocharger on the same ALCO 16-251 C engine. The turbocharger HS 5800 from Hedemora Turbo & Diesel has a max pressure ratio of 4.5, a global efficiency of 74%, and can be used on engines with rated power between 1500 and 3700 kW [45]. Before and after the update, the test engine was installed on a dedicated test bench (Figure 2) at IZAR facilities in San Fernando, Cadiz, Spain, (actually Navantia shipyards) to verify its performance. Comparative experimental data are provided in

Table 3. Experimental data.

Original Turbocharger ALCO 720
N [rpm]	800	900	1000	N [rpm]	800	900	1000
P2 [bar]	1.3	1.63	2.2	T4 [K]	772	813	817
T2 [K]	339.2	366.4	404.9	T3-T4 [K]	59	82	112
T2′ [K]	340.2	353.2	373.1	ΔP/P [%]	9.2	12.3	13.6
P3 [bar]	1.18	1.43	1.9	Pe [kW]	949	1493	2022
T3 [K]	831	895	929	BSFC [g/kWh]	231.744	213.639	208.231
New Turbocharger Hedemora HS5800
N [rpm]	800	900	1000	N [rpm]	800	900	1000
P2 [bar]	1.32	1.82	2.5	T4 [K]	745	751	715
T2 [K]	331.2	360.4	405.3	T3-T4 [K]	44	57	92
T2′ [K]	340.9	351.8	375	ΔP/P [%]	12.1	25.8	28
P3 [bar]	1.16	1.35	1.8	Pe [kW]	949	1503	2138
T3 [K]	789	808	807	BSFC [g/kWh]	231.744	212.176	196.918

Where: P2—Air pressure at the compressor outlet, T4—Exhaust gas temperature at turbo outlet, T2—Air temperature at the compressor outlet, Pe—Engine power, T2′—Air temperature at the engine inlet, N—Engine speed, P3—Exhaust gas pressure at the turbo inlet, ΔP/P—Engine cylinder sweep, T3—Exhaust gas temperature at the turbo inlet, BSFC—Brake-specific fuel consumption.

.

The experimental data recorded on the engine equipped with the Hedemora turbocharger were utilised to calibrate an engine model created and developed in the AVL Boost v 2022R1 tool, Figure 3, using the combustion mode Wiebe 2 Zones. This combustion mode considers two regions in the combustion chamber (burned and unburned zone) [46,47,48]. The main components of the engine model from Figure 3 are E1—engine; TC1—turbocharger; CO1—charge air cooler; PL1—intake manifold, C1…C16—engine cylinders; SB1, SB2 and SB3—system boundaries, 1…54—intake/exhaust pipe, MP1…MP37—measuring points, J1…J16—pipe connections, I1—Methanol injector.

In the first part of this study, for the engine model calibration, was used the simplified model of the turbocharger. The simplified model of the turbocharger is suitable for the analysis of the steady-state engine operation. This model does not consider the dynamics of the turbocharger and the efficiency of the turbocharger is kept constant during the engine cycle. For these initial simulations, the turbine layout calculation was considered, where the desired pressure ratio at the compressor is specified and the model adjusts the flow resistance of the turbine automatically until the energy balance over the turbocharger is satisfied.

Table 4. Relative deviations of the main parameters used for model calibration.

N	p2	Relative Deviation	T2	Relative Deviation	Pe	Relative Deviation	BSFC	Relative Deviation
[rpm]	[bar]	%	[K]	%	[kW]	%	[g/kWh]	%
800	1.33	0.5%	331.10	0.0%	951	0.3%	230.104	−0.7%
900	1.82	0.0%	367.20	1.9%	1507	0.2%	212.184	0.0%
1000	2.50	0.0%	405.50	0.1%	2142	0.2%	199.454	1.3%

shows the relative deviations concerning the experimental data of the main parameters considered for this initial model calibration.

For a better estimation of the common operating point of the engine–turbocharger unit, a global model of the turbocharger was integrated into the simulation model. The full model of the turbocharger requires as input the mechanical efficiency, the turbocharger rotor inertia, and the maps of the compressor and turbine. Using the specific characteristics of the turbocharger HS 5800 provided by Hedemora Turbo & Diesel, the compressor and turbine maps were built by the Turbocharger Map Generator of the AVL Boost tool.

The second part of this study was to assess the performance of the engine by fuelling it with biodiesel B20 (biofuel volumetric fraction of 20%, in diesel–biofuel blend). The third stage of the study was to investigate the impact on engine performance and pollutant emissions of using blends of diesel–methanol with methanol mass fractions of 10 and 20%, called DM10 and DM20. The final step of this study was to investigate the influence of fuelling the engine with tertiary blends of B20 and methanol using the same methanol mass fractions of 10 and 20%, called B20M10 and B20M20.

For the blended fuels, DM10, DM20, B20M10, and B20M20 as well as two fuelling conditions of the engine were considered: the direct injection of the blended original fuel with methanol inside the engine cylinders, and the dual-fuel mode with methanol injection outside the engine cylinders (fumigation solution) and the base fuel directly injected inside the engine cylinders. For the fumigation technique, the heat of evaporation of methanol was also considered. Figure 4 presents a group of windows with parameters selection for the Wiebe 2 Zone combustion model and the methanol injector conditions.

3. Results and Discussion

The main impact of using biofuels like biodiesel and methanol over the engine performance is the decrease in the brake effective power, caused by the decreased heating value of blended fuel when compared to that of conventional diesel fuel. As can be seen in Figure 5, in the case of using only B20, the reduction in brake effective power is of 3.4% for an engine speed of 800 rpm, 5.6% for 900 rpm, and 3.3% for 1000 rpm. The direct injection (DI) of methanol and conventional diesel fuel blends with 10 and 20% mass fractions of methanol leads to the reduction in power by 5.5% for 10% methanol fraction and by 11% for 20% methanol fraction at 800 rpm, 7.7% and 13.3% at 900 rpm, 5.4% and 10.8% at 1000 rpm (Figure 5a). For blends of biodiesel B20 and methanol in fractions of 10% and 20% methanol, the loss in brake effective power is of 8.5% for B20M10 and 13.7% for B20M20 at 800 rpm, 10.9% for B20M10 and 16.2% for B20M20 at 900 rpm, 8.4% for B20M10 and 13.5% for B20M20 at 1000 rpm. Considering the injection of methanol in the intake manifold after the charge air cooler (Figure 5b) for engine operation at 800 rpm and fuelled with DM10, the power loss is of 10.1 and 18.5% for DM20. Using the same blend at 900 rpm leads to a power loss of 13.5% for DM10 and 22.1% for DM20, with similar behaviour for an engine operating at 1000 rpm, 17.5% for DM10, and 22.1% for DM20. While replacing the conventional diesel fuel with B20 and keeping the same methanol fraction, the power loss at 800 rpm is of 12.8% for B20M10 and of 20.7% for B20M20, at 900 rpm it is of 16.3% for B20M10 and of 24.2% for B20M20, and at 1000 rpm it is of 20% for B20M10 and of 24.2% for B20M20 (Figure 5b).

Figure 6 shows the average peak fire pressure inside engine cylinders which is also influenced by the lower heating value of the fuel resulting in a similar behaviour as the brake effective power when changing the conventional diesel fuel with blends of diesel–methanol and biodiesel–methanol. Therefore, compared with the values registered for the conventional diesel fuel, the decrease in peak fire pressure while using the direct injection at 800 rpm for DM10 and DM20 is of 2.9% and 5.8%, at 900 rpm for DM10 and DM 20 it is of 4.4% and 8.7%, at 1000 rpm for DM10 and DM20 it is of 4.2% and 8.3%, respectively. The use of biofuel B20 also has an impact on the peak fire pressure, reducing it by 2% at 800 rpm, 3.2% at 900 rpm, and 2.9% at 1000 rpm. Blends of B20 and methanol will reduce the peak fire pressure even more, at 4.7% for B20M10 and 7.4% for B20M20 at 800 rpm, 7.2% for B20M10 and 11.2% for B20M20 at 900 rpm, 6.8% for B20M10 and 10.6% for B20M20 at 1000 rpm. In the case of the methanol fumigating solution, the evolution of the peak fire pressure displays important reductions of 8.9% for DM10, 11.8% for DM20, 10.5% for B20M10, and 13.2% for B20M20 at 800 rpm. At higher speeds, these reductions are more significant, at 12.2% for DM10, 17.4% for DM20, 14.6% for B20M10, and 19% for B20M20 at 900 rpm. At 1000 rpm, these mitigations are similar to those for DM10 at 21.1%, for DM20 at 29%, for B20M10 at 23%, and for B20D20 at 30.8%. Similar trends of the in-cylinder maximum pressure were reported in [41].

Since the fuel system of the engine contains individual pump injectors driven by the camshaft, the total mass of fuel injected was kept constant regardless of the blends of fuels used. The influence of using blends of diesel–methanol or biodiesel B20–methanol on the engine efficiency expressed by the brake-specific fuel consumption BSFC (Figure 7) is noticeable. BSFC increase can be justified by the lower heating value of the blend fuels used, which reduces the engine output. The increase in BSFC for a direct injection fuelling mode at 800 rpm is of 3.5% for B20, 5.8% for DM10, 12.3% for DM20, 9.3% for B20M10, and 15.9% for B20M20, compared to the original diesel fuel. At 900 rpm, this increases to 6% for B20, 8.3% for DM10, 15.4% for DM20, 12.2% for B20M10, and 19.3% for B20M20. At 1000 rpm, it seems that the higher turbulence intensity diminishes that growth to 3.4% for B20, 5.7% for DM10, 12.1% for DM20, 9.1% for B20M10, and 15.5% for B20M20 (Figure 7a). For the indirect injection of methanol, the increase in BSFC at 800 rpm is of 11.2% for DM10, 22.6% for DM20, 14.7% for B20M10, and 26.1% for B20M20, while at 900 rpm it is of 15.6% for DM10, 28.3% for DM20, 19.5% for B20M10, and 31.9% for B20M20, and at 1000 rpm it is of 13.5% for DM10, 28.3% for DM20, 16.9% for B20M10, and 31.9% for B20M20, compared with the results of the conventional diesel fuel. The results reported in [27] indicate an opposite trend of the BSFC but the results reported in [43] indicate a similar trend. The effect of indirect injection of methanol (fumigation) on BSFC is stronger in comparison with direct injection, being in good correlation with the effect on the engine brake output (Figure 7b).

The substitution of the original diesel fuel by mixtures of it with oxygenated fuels leads to lesser values of the stoichiometric air–fuel ratio for the blended fuels, conducting therefore to higher values of the relative air–fuel ratio (λ) compared to those of reference fuel condition operation (Figure 8). In the situation of fuelling only through the direct injection and operating the engine at 1000 rpm, the increase in the relative air–fuel ratio is of 11% for B20, 2.4% for DM10, 5% for DM20, 13.1% for B20M10, and 15.5% for B20M20. Operating the engine at 900 rpm, the increases in λ values are of 10.6% for B20, 1.9% for DM10, 4.1% for DM20, 12.3% for B20M10, and 14.1% for B20M20, and at the engine speed of 800, similar growths are registered at 12.9% for B20, 4.8% for DM10, 10.1% for DM20, 17.6% for B20M10, and 22.8% for B20M20 bigger (Figure 8a). In the case of methanol indirect injection (fumigation), the effect over the actual air–fuel ratio is slightly different because, when injected into the intake manifold, the airflow rate entering inside the combustion chamber diminishes, resulting in lower λ values compared with the case of direct injection. The relative air–fuel ratio for engine operation at 800 rpm and fumigation of methanol (Figure 8b) is increased by 4.9% for DM10, by 13.8% for DM20, by 17.9% for B20M10, and by 26.9% for B20M20. At 900 rpm, the relative air–fuel ratio is decreased by 0.1% for DM10, increased by 2.6% for DM20, increased by 10.3% for B20M10, and increased by 13.9% for B20M20. Operating at 1000 rpm, the λ values increase by 3.9% for DM10, decrease by 10.4% for DM20, increase by 15.4% for B20M10, and decrease by 1.4% for B20M20. This different behaviour at higher engine speeds for 10% and 20% methanol fractions could probably be explained by the occurrence of flow fluctuations in the induction stage caused by methanol indirect injection modifying the amount of air mass trapped inside the engine cylinders.

Figure 9 displays the variation in NOx emissions for the fuels and fuelling solutions investigated. In the case of fuels directly injected into the engine cylinders, the NOx emission will decrease as the methanol fraction increases. The cause of this evolution could be the variation in the cylinder charge peak fire temperature during combustion decreasing as the methanol fraction increases. Comparing the B20 with conventional diesel fuel, the NOx emissions are increasing, mainly due to the oxygen content existing in the B20 fuel. At 800 rpm, the increase in NOx is of 54.3%, at 900 rpm it is of 73.8%, and at 1000 rpm it is of 61.4%. The results concerning the addition of 10% methanol and comparing those obtained when fuelling with conventional diesel fuel, at 800 rpm, show a reduction of 19.1%, of 8.3% at 900 rpm, and of 10.4% at 1000 rpm, displaying the same trend of NOx pollutant as reported in [27,38,41]. Blending 10% in the mass fraction of methanol with B20 will lead to an increase in NOx of 27.9% for an engine operating at 800 rpm, of 58.3% at 900 rpm, and of 44.9% at 1000 rpm. For the higher fraction of 20% methanol blended with conventional diesel fuel, the NOx emissions drop by 36.2% at 800 rpm, by 17% at 900 rpm, and by 20.8% at 1000 rpm. The influence when blending the same methanol fraction with B20 consists in an increase in NOx of 2.3% at 800 rpm, of 41.8% at 900 rpm, and of 27.8% at 1000 rpm (Figure 9a). The fumigation of methanol in fractions of 10% and the use of conventional diesel fuel may lead to NOx reduction of 55.9% at 800 rpm, 43.7% at 900 rpm, and 61% at 1000 rpm. The same methanol fraction of 10% used along B20 reduces NOx emissions at 800 rpm and 1000 rpm by 21.5% and by 24.8%, respectively, while increasing them by 9.7% at 900 rpm. A fraction of methanol of 20% with conventional diesel fuel involves a reduction in NOx emission of 81.2% at 800 rpm, of 67.8% at 900 rpm, and of 56.7% at 1000 rpm. The cumulated effect on the NOx pollutants of fumigating a 20% fraction of methanol with B20 is to reduce these emissions by 63.6% at 800 rpm, 34.5% at 900 rpm, and 18.1% at 1000 rpm (Figure 9b). Reference [41] reported a similar dependence of NOx emissions by methanol fraction.

Figure 10 depicts the variations in peak fire temperature in combustion for the analysed conditions. As mentioned before, the peak fire temperature decreases with the increase in methanol mass fraction and the main cause for this can be explained by the reduced lower heating values of blended fuels compared to the conventional diesel fuel. Thus, for direct injection (Figure 10a) at 800 rpm engine speed, the temperature will decrease by 0.5% for B20, by 2.1% for DM10, by 4.2% for DM20, by 2.5% for B20M10, and by 4.6% for B20M20. At 900 rpm, the peak temperature will increase by 0.5% for B20, decrease by 0.9% for DM10, 1.8% for DM20, 0.4% for B20M10, and 1.5% for B20M20. For the engine running at 1000 rpm, it registered an increase in peak temperature by 0.3% for B20 but a reduction of 1% for DM10, of 2.1% for DM20, of 0.8% for B20M10, and of 1.9% for B20M20 (Figure 10a). The reduction in the peak fire temperature in the situation of fumigating methanol (Figure 10b) is as follows: for an engine running at 800 rpm, the reduction is of 6% for DM10, of 11.3% for DM20, of 6.4% for B20M10, and of 11.7% for B20M20; at 900 rpm, it is of 3.6% for DM10, of 7.1% for DM20, of 3.2% for B20M10, and of 7.3% for B20M20; at 1000 rpm, it is of 7.3% for DM10, of 2.7% for DM20, of 7.2% for B20M10, and of 2.4% for B20M20. At 1000 rpm for 20% methanol fraction induced by fumigation, the reduced λ values compared with those of the 10% methanol fraction could explain the slight increase in the maximum combustion temperature.

Figure 11 shows the evolutions of SOOT emissions. The mechanism of SOOT formations is an extremely complicated one. In this mechanism, the influence of oxygenated fuels over the SOOT formation plays an important part in SOOT reduction, mainly due to the oxygen content [49]. The presence of oxygen in the B20 fuel means better oxidation of carbon and hydrogen molecules, thus contributing to the reduction in SOOT emissions regardless of the methanol fuelling technology. These reductions of SOOT, in the case of fuelling the engine with B20, are at 800 rpm of 27.3%, at 900 rpm of 23.1%, and at 1000 rpm of 17.2%. The presence of methanol implies other factors which may reduce the oxidation of SOOT emissions, or even amplify the SOOT production [49]. Even if there is more oxygen available for the combustion process, because the peak fire temperature is lower, the oxidation of carbon and hydrogen molecules becomes slower, resulting in a higher quantity of SOOT emissions. For the direct injection of blended fuel of 10% methanol and conventional diesel fuel (Figure 11a), SOOT emissions increased by 0.2% at 800 rpm, 8% at 900 rpm, and 4.7% at 1000. The raising of the methanol fraction to 20% leads to an increase in SOOT emissions of 3.7% at 800 rpm, of 18.1% at 900 rpm, and of 10.7% at 1000 rpm. Evaluating the tertiary blend of biofuel, conventional diesel, and methanol, B20M10 indicates a reduction in SOOT emissions of 23.8% at 800 rpm, 15.7% at 900 rpm, and 12.3% at 1000 rpm, while for B20M10 the reduction is of 17.7% at 800 rpm, 6.3% at 900 rpm, and 6.1% at 1000 rpm. When injecting methanol into the engine intake manifold by fumigation (Figure 11b), less air remains trapped inside the engine cylinders, the cylinder charge suffers a local lack of air, and the mixture quality moves from a very lean to a slightly lean condition (Figure 8). In such circumstances at high temperatures, the fuel oxidation process is slowed down and a pyrolysis effect occurs, which massively contributes to the SOOT formation. Thus, the increase in SOOT emissions for DM10 at 800 rpm is of 240.6%, at 900 rpm it is of 280.5%, and at 1000 rpm it is of 315.9%. At the highest methanol fraction of 20%, DM20, there is a decrease in SOOT emissions of 14.5% at 800 rpm, 0.4% at 900 rpm, and an increase of 46% at 1000 rpm. Changing the conventional diesel fuel to B20 and using fumigation of methanol, the SOOT emissions increase by 174.5% for B20M10 at 800 rpm, 201.9% at 900 rpm, and 249.3% at 1000 rpm. For B20M20, there is a registered decrease in SOOT emissions at 800 rpm and 900 rpm by 26.1% and 18.1%, respectively, but an increase of 22.2% at 1000 rpm. This variation in SOOT, registering lower values at 20% methanol, could be correlated with the decrease in the air–fuel ratio mostly at 1000 rpm. However, the upper oxygen content existing in methanol and B20 normally favours fuel oxidation, but some reduction in the maximum combustion temperature could explain why, in the case of direct injection, the SOOT values are higher than those corresponding to the original diesel or B20 fuels. The results presented in this study concerning the use of biodiesel B20 and methanol blends indicate an opposite trend than those reported in [42,43].

4. Conclusions

The relevant conclusions of this study are summarised below:

• The main cause for the engine power loss is the lower heating value of the blends used as fuel. The effect of the blends of diesel–methanol and B20–methanol is similar on the peak fire pressure. This is valid for both methods of methanol supplying, with a slightly increased loss when fumigating methanol;
• The air–fuel ratio experienced growth, thus resulting in leaner mixtures inside the cylinder. In the case of direct injection of blended fuels, the behaviour is as explained above. In the case of the fumigation approach of methanol, the results indicate values lower or close to the reference conditions;
• NOx emissions are reduced as the methanol fraction is increased, and this behaviour is valid for both methods of methanol fuelling;
• When 10% methanol is supplied by the fumigation solution, regardless of the base fuel, either the original diesel or B20 the SOOT emissions are substantially increased. In the case of the 20% methanol fraction, the results indicate that SOOT emissions have an opposite trend, and this could be explained by the higher fuel quantity inside the cylinder as a homogeneous mixture before injecting the conventional diesel fuel or B20, which can lead to a better oxidation of carbon and hydrogen molecules similar to the RCCI technology;
• Evaluating both methods of methanol supply, the direct injection of blended methanol with diesel or B20, and the injection of methanol in the intake manifold (fumigation), the results indicate that the direct injection of blended fuels is a better option regarding the engine performance parameters and emissions;
• Considering an acceptable compromise of 10% in power loss, two fuels can represent an option for replacing the conventional diesel fuel. The first one is the diesel–methanol blend DM10 with a 10% mass fraction. Using this type of fuel, in direct injection technology, the NOx emission can be reduced up to 19% but with a compromise of 8% increase in SOOT emissions. The second one is B20–methanol 10% B20M10 with the same direct injection method. Using this blend as an alternative fuel, an increase in NOx of up to 58% with the benefit of reducing SOOT emissions by up to 23% could be obtained;
• All these theoretical results should be considered under the reserve of the hypotheses used for the simulations and must be confirmed by the experimental research.