Study on the Cumulative Effects of Using a High-Efficiency Turbocharger and Biodiesel B20 Fuelling on Performance and Emissions of a Large Marine Diesel Engine

Abstract

The marine sector represents probably the most powerful segment of international transport. Most ships use diesel engines for propulsion. Pollutant emission regulations with their continuous decline of acceptable limits put huge pressure on engine manufacturers. The use of low-quality fuels makes the marine sector a significant contributor to global pollution. The present study shows how turbocharger operating parameters and replacing diesel fuel with biodiesel B20 (20% oil and 80% diesel volumetric fractions) affect the performance, efficiency and pollutant emissions of a four-stroke diesel engine ALCO V16 251F for marine application. A combustion model developed with the AVL BOOST software was used to perform calculations using diesel fuel and biodiesel B20 for different turbocharger characteristics and injection timings. The model was calibrated against experimental data measured on a tested engine at the application site using diesel fuel and operating in a stationary condition of full load at 600, 700 and 800 rpm engine speeds. The results show that the cumulative effects of using an improved turbocharger associated with B20 fuelling under optimized injection timings could provide reductions of 45% for soot and 5% for NO_x, while maintaining the same engine performance obtained with diesel fuel operation.

1. Introduction

Maritime transport is one of the most powerful vectors for economic development because it represents the main axis of international trade (more than 70% of freight transport is made by the sea with an anticipated growth of 3.4% per year until 2050), contributing to 2.5% of worldwide greenhouse gas emissions [1]. Nowadays, the globalization process and the shift of substantial world industrial production to Asian countries have reached unprecedented levels in human history. In April 2018, the International Maritime Organization (IMO) issued Resolution MEPC.304(72) having three sustainability-oriented goals for the entire maritime shipping industry: (1) reduction of carbon compound (i.e., oxides and dioxides) emissions from new ships by the implementation of successive phases of the Energy Efficiency Design Index (EEDI), (2) reduction of carbon compound emissions in shipping by at least 40% by 2030, with attempts to achieve 70% reduction by 2050 (i.e., from the baseline year of 2008), and (3) reduction of GHG emissions in maritime shipping by at least 50% by 2050 with forced actions towards their complete elimination [2]. The IMO has also decided that to comply with the new emission targets, the energy efficiency of ships should increase by 40% by 2030 (i.e., compared to 2008 levels) and by 50–70% by 2050 [1,3,4].

Most marine ships use diesel engines. Since the new Tier III regulations adopted in 2016 for NO_x emissions relative to the previous Tier II emission standards involved a drastic reduction of NO_x and soot, the research on these topics became a major activity for engine manufacturers.

One possible method of increasing the efficiency of diesel engines and reducing pollutant emissions of NO_x and soot is the improvement of the turbocharging system characteristics. Nowadays, the turbocharging method is quasi-generalized as the turbocharging technique applied for large and very large diesel marine engines. A turbocharging system in its usual configuration is commonly made of a centrifugal compressor and an axial turbine with fixed geometry. Various methods of improving the turbocharger characteristics were studied, tested, and applied, each of them with different improvement degrees.

Several experimental and numerical studies show that a possible solution to improving a classical turbocharger with fixed geometry is to equip the turbocharger with a Variable Geometry Turbine (VGT) [5,6]. A VGT is commonly found as a turbocharger with a Variable Nozzles Turbine (VNT). At the same time as the VNT, a modern approach is to upgrade the compressor with Variable Diffuser Geometry (VDG) and/or Inlet Guided Vane (IGV). The benefits of a compressor provided with VDG and IGV are improved performance and efficiency: the reduction in turbocharger lag, lower emissions, higher fuel efficiency and extended operating range [6,7]. Using a system with VNT in combination with an Exhaust Gas Recirculation system (EGR), engine pollutant emissions can be reduced, and fuel efficiency can be increased. One of the main advantages of the VNT system is the ability to tilt the turbine nozzles and to operate the machine in an optimal condition, which can be found in reducing smoke emissions in the low engine speed area. Adjusting the nozzles angle results in lower exhaust back-pressure, reducing pumping loss and fuel consumption [8,9]. All methods mentioned above are usually used on small and medium size diesel engines, and they can also be considered for implementation on large diesel engines, along with the improved axial turbines that are usually used for power generation.

Another long-time studied solution is the use of tandem-bladed centrifugal compressors, which proved to have better performance than the conventional ones. Even if the research on tandem centrifugal impellers was started many decades ago, the results were not always promising. In 1977, an experimental investigation of a tandem-bladed centrifugal compressor observed little performance improvements and reduced damping from the impeller blading [10]. In 1991, a tandem-bladed compressor was patented, and it was found to have enhanced operating characteristics for surge, boost pressure and efficiency, but the improvement in efficiency was only 2% [11]. Nowadays, with the development of improved computer-aided design and calculation, new materials and manufacturing technologies, the tandem-bladed compressor demonstrated significant improvement on the conventional centrifugal compressors. For example, an experimental and numerical study [12] showed that a maximum of a 25% increase in the operation range relative to one of their examined geometrical variants could be observed.

Following the same idea, different researchers tried to manipulate the flow at the inlet of the compressor, by various fixed swirling devices [13,14] or by reducing the effective maximum peripheral velocity of the air impacting the front edges of the compressor blades [15]. All of them demonstrated important advantages regarding the surge effect and the compressor’s efficiency: the modification of the inlet velocity triangle affects the compressor power absorbed from the turbocharger shaft and therefore turbocharger speed can be either increased or decreased by changing the time-to-torque characteristics (turbo lag) [16].

Looking for high-efficiency turbochargers, another interesting approach is the optimization of the classical design process by modern numerical simulations and improved methods of calculation. Alongside the already well-known CFD methods, for example, one study [17] found a new method to improve the compressor map approach in engine performance modelling and simulation by characterizing enthalpy rise through the compressor. After removing data points likely influenced by heat transfer from the turbine to the compressor concurrently with estimating impeller outlet conditions using simplified geometry assumptions and a modified definition for compressor stage reaction, prediction errors for the rotation speed of the turbocharger were drastically reduced.

In addition to the numerical methods and the dimensional/structural design modifications, another way to modify the turbocharger characteristics is to change the fluids that are responsible for the working of the turbocharger. A large-scale implemented method is the Exhaust Gas Recirculation (EGR), which inputs the inlet of the compressor with a new air substitute gas, recirculating a portion of an engine’s exhaust gas back to the engine cylinders through the turbocharger, with the scope of reduction of the nitrogen oxide (NO_x) emissions. Various extensive studies [16] were conducted to examine the working of the turbocharger in these conditions. Different materials or coatings for the compressor are needed due to the new environment [18,19]. Moreover, changing the fluid at the turbine inlet will drastically influence the turbocharger attributes. The burned gases in this area are coming from the Internal Combustion Engine (ICE) and they are directly dependent on the type of fuel involved in the combustion process.

The type and the composition of the fuel involved in the chemical combustion reactions significantly influence the combustion characteristics and consequently the generation of pollutant emissions. Therefore, obtaining new environmentally friendly and renewable alternative fuels is a constant and common pursuit of fuel producers and engine manufacturers.

A first step in fulfilling this purpose is the use of alternative fuels such as biofuels and alcohols, which are produced from renewable sources, demonstrating lower CO₂ emissions over the entire fuel lifecycle. Biodiesel fuel types are mixtures in different fractions of renewable biofuel with petroleum-derived diesel fuels as a natural and sustainable energy resource. When mixing fuels, it is important to ensure that components of blended fuel are compatible with each other; if not, issues can arise because the mixed fuel can lose its sedimentation stability due to the asphaltenes sediment formation. However, these matters are encountered more in heavy residual marine fuels [20] With minimal aromatic hydrocarbons and sulphur content, the biodiesel blends do not normally face this kind of problem. Furthermore, they are characterized by very good lubricity, high cetane number and elevated flash point. Biodiesel fuel blends such as B7 and B10 are already used in small and medium diesel engines without the need for substantial modifications of the fuelling system but with dedicated adjustments which are required for original performance recovery. Most of these engines are supercharged engines. However, in the last decades, new research has focused on the possibility of expanding the use of biodiesel fuels with higher biofuel contents such as B20, B30, B50 and even B100 [1,21,22,23,24,25,26].

In this sense, many studies have investigated the environmental impact of different concentrations of biodiesel used as the main fuel in diesel engine operation. The results obtained until now show significantly reduced values of CO-emissions, sulphur level, unburned hydrocarbons and particulate matters in the exhaust gases in comparison with those of conventional diesel fuel [27,28,29,30].

In the field of the medium, large and very large marine diesel engines, mainly fuelled by diesel or even by heavy fuel oil, the use of biodiesel fuels seems to be a promising path to be followed as an important way to reduce greenhouse gas emissions [29]. These research studies are in progress, showing that high fractions of biofuel (up to 30%) can be used without major adjustments and changes to the original engine configurations. For the moment, the standards for marine diesel fuels do not include any percentage of biodiesel in their composition [31].

This study is focused on the possibilities of replacing the conventional diesel fuel for marine diesel engines with biodiesel B20 (20% volumetric fraction of biofuel mixed with diesel fuel) and the cumulative effects of improving the turbocharger characteristics with the use of B20, highlighting the results on engine performance and pollutant emissions.

The novelty of the present research work consists of emphasizing what could be the influence of biodiesel B20 fuelling on the combustion process, performance, and emissions of a large marine diesel engine when this fuelling mode is associated with an improvement of the engine turbocharging system and when a slight adjustment in injection characteristics is accomplished.

2. Materials and Methods

The technical specifications of the diesel marine engine used for this study are shown in

Table 1. Engine specifications.

Application	Marine
Brake power	3900 HP
Rated speed	1100 rpm
BSFC@1100 rpm	224.7 g/kWh
BMEP	17.93 bar
Bore	228.6 mm
Stroke	266.7 mm
Configuration	V16
Displacement	175.1 L
Compression ratio	11.5
Valve timing	Fixed, Camshaft driven
Valve overlap	140°
Injection System	Mechanical, pump-injector system
Injector	9 holes × 0.375 mm
Injection pressure	260 bar

.

The engine was provided with its original turbocharger delivered by the engine manufacturer. A revamped and improved turbocharger (compression pressure ratio elevated by 0.2 and compressor efficiency by 0.5%) dedicated to the same engine for marine application was also considered in this study for the simulation stage. The compressor map is presented in Figure 1, along with the compressor operating lines before and after its revamping.

For the separate testing of the turbocharger, a specific test rig was designed and executed. The central focus was to study the main characteristics and the parameters for different operating regimes of the tested turbochargers. A layout of the test rig is displayed in Figure 2. The test bench is like the real working principle, with the main difference consisting of a replacement of the internal combustion engine with an independently controlled hot gas generator. A supplementary air source adds air to the circuit, especially in the starting phase when the air from the compressor is still low. An excess air discharge circuit was provided in the test rig to be able to control the compressor output pressure and avoid the compressor surge effect.

In addition to the parameters to be monitored for the control of the auxiliary circuits (water, oil, and fuel), there are four main points of interest: 1. the compressor air inlet, 2. the compressor air outlet, 3. the turbine hot gases inlet, and 4. the turbine outlet. For all these points, the measurements recorded the temperature and pressure. Moreover, two simultaneous separated probes indicated the turbocharger speed.

Based on these measured parameters, various special protections and automatic limitations were considered, to eliminate the possibility of human error and to protect the turbocharger. Special attention and clear protection limits were established regarding the vibrations (axial, radial–vertical and radial–horizontal). The protections and red lines were settled according to the maximum parameters specified in the manufacturer’s manual.

The measured data on the real engine operation as a marine application were used to calibrate a model developed in the AVL BOOST tool (Figure 3) using the combustion mode expressed by AVL MCC (Mixing Controlled Combustion) [32] (Figure 4). Data used for calibration of the model, such as thermodynamic parameters from compressor inlet and outlet, turbine intake and outlet, and turbine speed, were previously recorded on the turbocharger test bench and validated at the application site. The engine speed was also recorded. An AVL MCC model is dedicated to direct injection engines and allows a detailed configuration of the fuel injection timing. The injection timing and fuel injected quantity were those specified by the manufacturer. NO_x formation was calculated based on the extended Zeldovich mechanism coupled with three other additional reactions for N₂O creation. Soot estimation was achieved with a mechanism involving two reactions of formation and oxidation which are based on chemical kinetics. The main components of the model from Figure 2 are the following: E1.engine; TC1, turbocharger; CO1, charge air cooler; PL1, intake manifold, C1…C16, engine cylinders; SB1, SB2 and SB3, system boundaries, 1…53 intake/exhaust pipe, MP1…MP37 measuring points, J1…J16 junctions. Engine model validation was conducted according to the effective performance data from the technical specifications provided by the engine manufacturer and the experimental data measured at the application location. All the experimental data were gathered after the turbocharger was refurbished and tested on the test bench.

AVL MCC combustion model theory [32]

The model considers the effects of the premixed (PMC) and diffusion (MCC) controlled combustion processes according to Equation (1):

$$\frac{d{Q}_{total}}{d\alpha }=\frac{d{Q}_{MCC}}{d\alpha }+\frac{d{Q}_{PMC}}{d\alpha }$$

(1)

MCC or mixing controlled combustion:

In this regime, the heat release is a function of the fuel quantity available (f₁) and the turbulent kinetic energy density (f₂):

$$\frac{d{Q}_{MCC}}{d\alpha }=C_{Comb}\ast f_1\left(m_f, Q_{MCC}\right)\ast f_2\left(k,V\right)$$

(2)

with

$$f_1\left(m_f,Q\right)=\left(m_f-\frac{Q_{MCC}}{LCV}\right)\ast {\left(w_{oxygen,available}\right)}^{C_{EGR}}$$

(3)

$$f_2\left(k,V\right)=C_{Rate}\ast \frac{\sqrt{k}}{\sqrt[3]{V}}$$

(4)

where

$Q_{MCC}$	cumulative heat release for the mixture controlled combustion [kJ]
$C_{Comb}$	combustion constant [kJ/kg/deg CA]
$C_{Rate}$	mixing rate constant [s]
$k$	local density of turbulent kinetic energy [m2/s2]
$m_F$	vaporized fuel mass (actual) [kg]
$LCV$	lower heating value [kJ/kg]
$V$	cylinder volume [m3]
$\alpha$	crank angle [deg CA]
$w_{Oxygen,available}$	mass fraction of available oxygen (aspirated and in EGR) at SOI [-]
$C_{EGR}$	EGR influence constant [-]

Conservation equation for the kinetic energy of the fuel jet:

Since the distribution of squish and swirl to the kinetic energy are relatively small, only the kinetic energy input from the fuel spray is taken into account. The amount of kinetic energy imparted to the cylinder charge is determined by the injection rate (first term on RHS). The dissipation is considered as proportional to the kinetic energy (second term on RHS) giving:

for ‘Revised’ TKE calculation:

$$\frac{d{E}_{kin}}{dt}=0.5\ast C_{turb}\ast {\dot{m}}_F\ast v_F^2-C_{Diss}\ast E_{kin}^{1.5}$$

(5)

$$k=\frac{E_{kin}}{m_{F,1}\left(1+{\lambda }_{Diff}{m}_{stoich}\right)}$$

(6)

for ‘Default’ TKE calculation (this is an older status of the model):

$$\frac{d{E}_{kin}}{dt}=0.5\ast {\dot{m}}_F\ast v_F^2-C_{Diss}\ast E_{kin}$$

(7)

$$k=\frac{C_{turb}\ast E_{kin}}{m_{F,1}\left(1+{\lambda }_{Diff}{m}_{stoich}\right)}$$

(8)

$$v=\frac{{\dot{m}}_F}{{\rho }_F\ast \mu A}$$

(9)

where

$E_{kin}$	kinetic jet energy [J]
$C_{turb}$	turbulent energy production constant [-]
$C_{Diss}$	dissipation constant; ‘Revised’: [J−0.5/s]; ‘Default’: [1/s]
${\dot{m}}_{F,1}$	injected fuel mass (actual) [kg]
$v$	injection velocity [m/s]
$\mu A$	effective nozzle hole area [m2]
${\rho }_F$	fuel density [kg/m3]
$n$	engine speed [rpm]
$m_{stoich}$	stoichiometric mass of fresh charge [kg/kg]
${\lambda }_{Diff}$	air excess ratio for diffusion burning [-]
$t$	time [s]

The first step of the study was to simulate the engine operation when fuelled with 100% diesel fuel to calibrate the model and create reference points for three engine operating speeds, 600, 700 rpm and 800 rpm, at full load conditions. The chosen engine operating speeds allowed experimental data to be recorded. The second step of the study involved the simulation of the engine operation when fuelled with a mixture of diesel and biofuel in a volumetric fraction of 20% named biodiesel B20. The third step in the simulation study was the assessment of the cumulated effect of the improved turbocharger characteristics associated with the use of B20 fuelling. The fourth step in this study represented the simulation of the cumulated effect of using the improved turbocharger when fuelling the engine with B20 and optimized injection timing by delaying the fuel injection with 5, 7, 9 and 11 CA degrees. The fuel injection parameters are in

Table 2. Fuel injection parameters.

Engine Speed	BSOI	TSOI	TEOI	BEOI
600	−14	−13	−1	0
700	−14	−13	−1	0
800	−15	−14	2	3

, where BSOI is bottom start of injection, TSOI is top start of injection, TEOI is top end of injection, and BEOI is bottom end of injection.

In all these cases, the amount of fuel injected was kept the same as the original one depending on the engine speed. The properties of the fuels used for this study and those of biofuel are shown in

Table 3. Physico-chemical properties of different fuels [21,22,23,24,25,26,27,28,29,30,33,34,35,36,37].

Properties	Diesel Fuel	Biofuel	Biodiesel B20
Chemical formula	C12H26-C14H30	C19H36O2	C15H23O
Molecular weight (g/mol)	170–220	292.6	219
Density @ 20 °C (kg/m3)	810–880	887	857.4
Boiling Point (°C)	125i–400f	330	200
Viscosity (20 °C) (cSt)	3.35	8.06	5.12
Flash Point (°C)	65–88	140	85
Autoignition Temperature (°C)	204–340	380	-
Cetane number	40–55	55–56	52.5
Air/Fuel ratio at Stoichiometric	14.7	12.6	14.2
Lower heating value (MJ/kg)	42.5	38.8	40.5
Heat of vaporization (kJ/kg)	260	350	277
Carbon content (%wt)	87	76.9	82.2
Hydrogen content (%wt)	12.6	12.4	10.5
Oxygen content (%wt)	0.004	10.7	7.3
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	-	-

Fuel data was used in the internal database of the simulation code; i—initial boiling point; f—final boiling point.

.

3. Results and Discussions

Fuelling a diesel engine with biofuels or blends of conventional diesel fuel can be beneficial for overall engine performance regarding soot emissions but negative regarding power(Pe) and torque(Me) performance. Usually, these effects are caused by the reduced lower heating value of the biodiesel compared to the lower heating value of the original diesel fuel. In Figure 5a,b, the reduction in torque and power can be seen when comparing the fuelling with original diesel fuel and biodiesel B20. Furthermore, considering the operation with a more efficient turbocharger, the effects are superior on the engine performance. Using biodiesel B20 leads to a loss in torque and power up to 0.8% at 600 rpm and 700 rpm and 0.3% at 800 rpm, compared to those for diesel fuel, with similar results being found in [27,28,37]. Considering the engine operation with a turbocharger having improved efficiency, an increase in torque and power by 2.1% for 600 rpm, 1.6% for 700 rpm and 1.2% for 800 rpm compared to diesel fuel results. Coupling the improved turbocharger with B20 fuelling again provides an increase in torque and power by 1.5% for 600 rpm, 1% for 700 rpm and 1.1% for 800 rpm compared to the use of classical diesel fuel reference points. In these figures 600 D stands for engine running at 600 rpm using only Diesel fuel, 600 D 0.2 stands for engine running at 600 rpm using Diesel fuel and a compressor compressing ration increased by 0.2, 600 B20 stands for engine running at 600 rpm using biodiesel B20 fuel and 600 B20 0.2 stands for engine running at 600 rpm using biodiesel B20 fuel and a compressor compressing ratio increased by 0.2. This applies for 700 rpm and 800 rpm.

The influence of the alternative fuel, biodiesel B20, on brake-specific fuel consumption (BSFC) is somewhat negative. For an engine operating with B20, Figure 6a,b depict an increase in BSFC up to 0.7% for 600 rpm, 0.8% for 700 rpm and 0.3% for 800 rpm, compared with diesel fuel results, and this behaviour is associated with some loss of torque and power performance; a similar effect of using B20 was confirmed in [27,28,37]. When the improved turbocharger is considered without the use of the alternative fuel, there is a positive effect on the BSFC, leading to a reduction of 2% at 600 rpm, 1.6% at 700 rpm and 1.2% at 800 rpm, compared to the reference conditions, while using B20 with the more efficient turbocharger may lead to a reduction of BSFC by 1.5% at 600 rpm, 1% at 700 rpm, and 1.1% at 800 rpm.

The effect on the relative air–fuel ratio (λ) of replacing diesel fuel with B20 shows an increase of around 14% for the three engine speeds considered (see Figure 5b). The main cause of this increase is the higher content of oxygen existing in the biodiesel fuel. The relative air–fuel ratio also increases as a direct effect of a more efficient turbocharger by 19.6% for 600 rpm, 17.2% for 700 rpm and 13.4% for 800 rpm for pure diesel fuel. Coupling the fuelling of the engine with B20 with the operation of a more efficient turbocharger, the relative air–fuel ratio increases by 36.1% for 600 rpm, 33.4% for 700 rpm and 29% for 800 rpm. Clearly, the reasons for these increases in the air–fuel ratio are the extra air quantity delivered by the more efficient compressor and the high content of oxygen existing in the B20 fuel.

Soot and NO_x emissions are presented in Figure 7a,b. The main benefit of using oxygenated fuels, in this case B20, as alternatives to conventional diesel fuel, is the reduction in soot emissions. When fuelling the engine with B20, there is a drop in soot emission up to 35.3% for operating at 600 rpm, 36.9% at 700 rpm and 40.8% at 800 rpm. The soot reduction is in turn associated with an increase in NO_x emissions, by 42.6% for 600 rpm, 44.2% for 700 rpm and 41.5% for 800 rpm, compared with pure diesel operation, which represents a real drawback. Similar results have been reported in [21,22,28,37]. Positive effects come by using a more efficient turbocharger when soot emissions are decreasing by 32.6% for 600 rpm, 29.9% for 700 rpm and 29% for 800 rpm, while a decrease of NO_x emission by 13.3% for 600 rpm, 11.9% for 700 rpm and 6.3% for 800 rpm is also registered. Fuelling the engine with B20 coupled with the use of a more efficient turbocharger involves a reduction in soot emission by 53.1% for 600 rpm, 52.9% for 700 rpm and 55.1% for 800 rpm, while the NO_x emission is roughly increased by 21% for the three operating speeds considered, relative to the reference diesel results. The main cause for soot reduction in all cases is the higher oxygen content in the combustion chamber which ensures better oxidation of the existing fuel. It can be noticed that NO_x emissions are increasing only when using the B20 fuel, mainly due to the elevated oxygen content leading to better conditions for the NO_x formation mechanism.

Figure 8a indicate that the in-cylinder maximum pressure has a small variation when the engine is fuelled with B20 compared to the reference, mainly because conventional diesel fuel and B20 have similar combustion properties. When considering the engine operation with a more efficient turbocharger, there is an increase in the in-cylinder maximum pressure by 9.1% at 600 rpm, 8.1% at 700 rpm and 6.7% at 800 rpm. Together, a more efficient turbocharger and the engine fuelling with B20 lead to a rise of the in-cylinder maximum pressure by 11.6% for 600 rpm, 10.5% for 700 rpm and 9.2% for 800 rpm. Figure 8b reveals that the peak pressure rise rate is insignificantly affected by changing the conventional diesel fuel to biodiesel B20. This behaviour is relatively normal because the conventional diesel fuel and biodiesel B20 have fairly similar combustion properties.

The substitution of the diesel fuel and considering a more efficient turbocharger will have a significant impact on the combustion process inside the cylinder. From Figure 9a, it appears that changing the diesel fuel with B20 involves a shortening of the initial phase of the combustion, defined as the difference between the angle where 10% of the fuel was burned and the angle of start of combustion, further called Δα_i, by 2.9% for 600 rpm, 2.7% for 700 rpm and 2.2% for 800 rpm. Similar behaviour occurs in the main phase of combustion, defined as the difference between the angle where 90% of the fuel was burned and the angle where 10% of the fuel was burned, further called Δα_m, (Figure 9b). The reductions are more significant, being by 27.3% at 600 rpm, 27.4% at 700 rpm and 29.8% at 800 rpm. When considering the usage of a more efficient turbocharger, the variation of the initial combustion phase has the same behaviour as when diesel fuel was replaced by B20. The reduction in the initial phase of the combustion is 5.8% for 600 rpm, 5.6% for 700 rpm and 3.8% for 800 rpm, while the increase in the main phase of the combustion is 12.9% for 600 rpm, 11.2% for 700 rpm and 7.7% for 800 rpm. These reductions in the initial stage of combustion for a high-efficiency turbocharger are related to the higher pressure and temperature conditions which accelerate the kinetics of the chemical reactions and consequently induce the shortening of the premixed combustion independently of engine speed. The changing of the pure diesel fuel with biodiesel B20 also diminishes the premixed combustion duration because of the increased biodiesel reactivity. While changing the pure diesel fuel with B20, the main phase of combustion behaves similarly to the premixed combustion phase, but using a more efficient turbocharger regardless of the fuel used the variation of the main combustion phase instead has a reverse behaviour. This could be explained by the substantial leaning of the cylinder charge which occurs at high air excess and reduces the flame propagation speed, emphasizing thus the dominant effect of the air fuel mixture.

Considering both solutions of fuelling the engine with B20 and using the improved turbocharger, the effects on the initial combustion phase are reductions of 8.7% at 600 rpm, 7.5% at 700 rpm and 6.4% at 800 rpm. On the main combustion phase, the reductions are 19.5% at 600 rpm, 21% at 700 rpm and 25% at 800 rpm relative to the initial diesel fuel reference conditions.

A possibility to reduce the growth of NO_x emission following the fuelling with B20 is to retard the injection timing. By doing so, there may be an engine derating present, due to the fact that retarding the injection will result in a shift of the maximum pressure peak to the expansion stroke and a reduction of it. Studying multiple values of injection retarding while considering a more efficient turbocharger and fuelling with B20, NO_x emission may increase by 2.4% for 600 rpm, 5.4% for 700 rpm and 9.3% for 800, while the reduction in soot is 43.7% for 600 rpm, 42.9% for 700 rpm and 46.2% for 800 rpm, compared with the reference results. The above results are obtained retarding the injection by 7 degrees to the expansion stroke (Figure 10a,b and Figure 11). It has been found that retarding the injection by 9 or 11 degrees may lead to an increased engine derating, compared with retarding by 7 degrees results where the engine derating is absent or minimal. In the situation where the injection timing is retarded by 11 degrees, there is a reduction in both soot and NO_x emission with a slight degrading in the engine’s power of up to 3%. For 600 rpm, soot emissions drop by 35.5% at 600 rpm, by 37.8% at 700 rpm and at 800 rpm by 39.4%, while NO_x emissions dropped by 8.8% for 600 rpm, 5.8% for 700 rpm and 2.1% for 800 rpm.

Figure 11 summarises the effect of altering the injection timing from the original values to that modified by 11 CA degrees on the engine brake power. It can be noted that by delaying the injection timing, the engine brake power drops up to 3%.

4. Conclusions

The conclusions of this study can be summarized as follows:

-

The engine operation with a high-efficiency turbocharger having improved efficiency by 0.5% and higher compression pressure ratio by 0.2 relative to the original turbocharger specifications enhances the engine performance and efficiency by around 1.5% regardless of the fuel used for engine fuelling,

-

Soot and NO_x emissions are reduced by around 30% and 10% regardless of the fuel used for engine operation,

-

The substitution of the classical diesel fuel with biodiesel B20 mitigates the engine output and efficiency by around 0.8% for the original turbocharger operation, while the soot emissions are reduced by 37% and NO_x increased by 43%,

-

The cumulative effects of using a high-efficiency turbocharger and biodiesel B20 for the original engine adjustments of injection timings are related to the engine performance and efficiency increasing by 1.2%, soot emissions reduction by 53% and NO_x emissions increasing by 21%,

-

At the optimized injection timing, the cumulative effects are still an important reduction of soot by 45% and a small increase of NO_x by 5% for practically the same engine performance obtained on diesel fuel operation,

-

When a slight derating of 3% for the engine output is accepted, then a simultaneous reduction of soot by 37% and NO_x by 6% could be reached.

All the results have been obtained with the following simplifying assumptions:

▪

The properties of B20 were calculated starting from the properties of the conventional diesel fuel and biofuel B100;

▪

The properties of B100 were considered for rapeseed oil;

▪

The amount of fuel injected was kept constant for all the simulation conditions;

▪

The retarding of fuel injection was not verified in the experimental activities where experimental data for model calibration was recorded.

As future work, the team will consider experimental activities at the engine test bench to validate the results.