Fischer-Tropsch diesel and biofuels exergy and energy analysis for low emissions vehicles

applied sciences Article Fischer-Tropsch Diesel and Biofuels Exergy and Energy Analysis for Low Emissions Vehicles Felipe Andrade Torres 1,2 , Omid Doustdar 3, Jose Martin Herreros 3, Runzhao Li 3, Robert Poku 3, Athanasios Tsolakis 3,*, Jorge Martins 4and Silvio A. B. Vieira de Melo 1,5   Citation: Torres, F.A.; Doustdar, O.; Herreros, J.M.; Li, R.; Poku, R.; Tsolakis, A.; Martins, J.; Vieira de Melo, S.A.B. Fischer-Tropsch Diesel and Biofuels Exergy and Energy Analysis for Low Emissions Vehicles. Appl. Sci. 2021,11, 5958. https:// doi.org/10.3390/app11135958 Academic Editor: Adrian Irimescu Received: 20 May 2021 Accepted: 24 June 2021 Published: 26 June 2021 Publisher’s Note: MDPI stays neutral with regard to jurisdictional claims in published maps and institutional affiliations. Copyright: © 2021 by the authors. Licensee MDPI, Basel, Switzerland. This article is an open access article distributed under the terms and conditions of the Creative Commons Attribution (CC BY) license (https:// creativecommons.org/licenses/by/ 4.0/). 1 Industrial Engineering Program, Polytechnic School, Federal University of Bahia, Salvador 40210-630, Brazil; [email protected] (F.A.T.); [email protected] (S.A.B.V.d.M.) 2Center of Exact and Technological Sciences, Department of Mechanical Systems, Federal University of Recôncavo of Bahia, Cruz das Almas 44380-000, Brazil 3Department of Mechanical Engineering, School of Engineering, University of Birmingham, Birmingham B15 2TT, UK; [email protected] (O.D.); [email protected] (J.M.H.); [email protected] (R.L.); [email protected] (R.P.) 4Department of Mechanical Engineering, University of Minho, 4800-058 Guimaraes, Portugal; [email protected] 5 Interdisciplinary Center in Energy and Environment, Federal University of Bahia, Salvador 40170-115, Brazil *Correspondence: [email protected] Abstract: This research investigates the effects of a synthetic diesel-like fuel (Fischer-Tropsch diesel) and biofuels (ethanol and biodiesel) fuel blends on the energy-exergy efficiencies and gaseous exhaust emissions characteristics of a compression ignition engine. Two blends of alternative fuels denoted as E15B35FTD50 (15% ethanol, 35% biodiesel, and 50% Fischer-Tropsch diesel) and E15B35D50 (15% ethanol, 35% biodiesel, and 50% diesel) were experimentally studied on a single-cylinder diesel engine and compared to diesel fuel. The results show that the energetic and the exergetic efficiencies of the alternative fuels are comparable to those of the engine fueled with diesel fuel. The unburnt HC, NO, N 2 O, and NH 3 emissions were reduced for the two alternative fuel blends compared to diesel, while CO emissions increased. The light HC species were found to slightly increase for the alternative fuel blends in comparison with diesel fuel. However, the total HC was considerably reduced by the combustion of E15B35FTD50 not only when compared to the diesel fuel combustion, but also when compared to E15B35D50. Overall, these results may contribute to identifying advantages and limitations in terms of energetic-exergetic analysis and emissions for the new generation of conventional diesel and hybrid electric vehicles that aim to achieve future emissions regulations. Keywords: biofuels; energy; exergy; emissions; ethanol; Fischer-Tropsch diesel; engine 1. Introduction Internal combustion engines (ICE) are the worldwide major powertrain system that enables road transportation. Although electric vehicles are gradually gaining inroads in the automotive market, ICE still continues to be widely used worldwide [ 1 , 2 ] both in ICE vehicles as well as in hybrid electric vehicles (HEV). Full electrification is still in the early stages of development in most countries. For example, in Brazil, hybrid and flexible-fuel vehicles are projected for the future with 52% and 32% contribution by 2050, respectively. The electrification and biofuel synergies are required to enhance Brazil’s capability for meeting the Paris Agreement targets [3]. Current legislation is forcing the reduction of exhaust emissions, particularly nitrogen oxides (NO x ) and particulate matter (PM) for compression ignition (CI) engines. One of the alternatives to reduce engine exhaust emissions towards this effort is the use of alternative fuels [ 4 , 5 ]. Biofuels and hybrid electric vehicle synergies can effectively contribute to mitigating the dependence on fossil fuels as well as the greenhouse gases and air quality issues associated with their emissions. These synergies are not only applicable to the Appl. Sci. 2021,11, 5958. https://doi.org/10.3390/app11135958 https://www.mdpi.com/journal/applsci Appl. Sci. 2021,11, 5958 2 of 21 utilization of the fuel in the vehicle but also in a well-to-wheel system evaluation. This reinforces the renewable energy proposal, especially in countries where the electrical matrix is primarily renewable, such as the case of Brazil. Fischer-Tropsch (F-T) fuels, as one of the common biofuels, are synthetic fuels that can be produced from catalytic conversion processes using biomass (biomass-to-liquid or BTL) as raw materials that are able to be used to synthesize diesel-like fuels [ 6 ], among others. Therefore, the F-T diesel might be considered a renewable fuel depending on the type of raw material that was used to convert it into synthetic diesel-like fuel [ 7 ]. However, not usually used in diesel engines, ethanol is the most produced biofuel in the world and is another key biofuel that can be used as a fuel component in CI engines [ 8 , 9 ]. Biodiesel is a diesel-like fuel that can partially substitute the diesel fuel in a diesel engine. These alternative fuels can form binary and ternary fuel blends to be used in CI engines without requiring major powertrain modifications [10]. Researchers have investigated the use of biofuels in CI engines. Emiro˘glu and ¸Sen [ 11 ] evaluated diesel/biodiesel/alcohol fuel blends on a single-cylinder diesel engine. Increases in nitrogen oxides (NO x ) and hydrocarbons (HC) emissions, decrease in smoke, and carbon monoxide (CO) were obtained compared to diesel fuel combustion. They also reported that the thermal (or energy) efficiencies of the engine were very similar when using biodiesel/alcohols or when using diesel fuel. Valencia Ochoa et al. [ 12 ] investigated two types of biodiesel blended with diesel fuel on a single-cylinder diesel engine. They reported that the blends had lower CO, HC, smoke, and carbon dioxide (CO 2 ), but the NO x emissions increased compared to diesel fuel engine operation. Additionally, the results showed that diesel fuel had the highest energy efficiency among the tested fuels. Choi et al. [ 13 ] assessed two blends of GTL (gas-to-liquid) and biodiesel on a single-cylinder diesel engine and compared them to pure GTL fuel. It is concluded that the GTL100 (100% GTL fuel) exhibited lower NO x , CO, HC, and soot than the biodiesel blended with GTL fuel. Venu et al. [ 14 ] evaluated a ternary blend of diesel/biodiesel/ethanol with and without alumina (Al 2 O 3 ) nanoparticles on a single-cylinder diesel engine. It is reported that the ternary blend decreased the CO, HC, smoke, and PM, although the NO x increased and the engine thermal efficiency was lower for the blend than for diesel fuel. Furthermore, studies showed a remarkable emission reduction with hybrid electric vehicles running with biofuels [ 15 ]. García et al. [ 16 ] investigated two advanced dual-fuel combustion modes using bioethanol as the main fuel for hybrid passenger vehicles, a pre-chamber ignition system (PCIS) using bioethanol and hydrogen, and reactivity-controlled compression ignition (RCCI) combustion mode fueled with bioethanol/diesel. As a result, the RCCI mode has shown the highest potential to decrease the NO x emissions while presenting the highest benefits in energy consumption, however had penalties in terms of CO 2 emissions. García et al. [ 17 ] also studied dual-fuel diesel and E85 (85% ethanol and 15% gasoline) in a series hybrid vehicle concept and concluded that ultra-low engine-out emissions could be achieved for NO x . Furthermore, regarding advanced combustion strategies, Kumar and Sandhu [ 18 ] recently reviewed the impact of partially premixed combustion (PPC) on CI engines and summarized that the integrated application of PPC mode in addition to high octane number fuels, such as ethanol and butanol, in modern CI engine is a promising key alternative to match the stringent emission legislation. Macián et al. [ 19 ] evaluated the potential of synthetic fuels towards the forthcoming EURO VII emissions limits. The researchers reported that the 2030 target to CO 2 emissions could be achieved in a well-to-wheel base as well as they reached ultra-low NO x emissions (<0.2 g/kWh) with zero-soot emissions through the engine operational map. However, they concluded that in order to reduce CO 2 emissions and reduce fuel consumption, vehicle hybridization must be considered. Besides, the automotive industry is continuously developing diesel combustion system designs to achieve the CO 2 targets and improve overall engine-out emissions [20]. Most of these works have evaluated the energetic analysis of alternative fuels in CI engines in terms of the First Law of Thermodynamics. This is the most frequently employed method, but solely it is not sufficient to establish the losses and the efficiency of a Appl. Sci. 2021,11, 5958 3 of 21 system [ 21 ]. Exergy analysis is a method that combines both the First and the Second Laws of Thermodynamics to determine the losses (or irreversibilities) of a system [ 22 ]. It helps in assessing the source of these losses and allows more detailed information regarding the efficiency of the engine. Besides, it is an important tool to provide the exergy destruction analysis of a system [23]. Table 1shows a summary of the results of energy and exergy analysis of previously conducted investigations. It becomes clear that the exergy efficiency is of great interest to the scientific community, as it directly identifies the existing destructions during the engine operation process. It is well established by the literature that increasing the engine load leads to an increase in the energy and exergy efficiencies [ 24 , 25 ], whilst increasing the engine speed causes the opposite effect [ 26 , 27 ]. Therefore, a detailed analysis of the effects of a specific fuel or blend on the energetic and exergetic efficiencies could be performed by comparing the efficiencies and the losses of the engine. Based on the aforementioned studies, this work contributes to the current literature by applying an energetic analysis (fuel energy, energy losses, outlet work, and energy efficiency) and exergetic analysis (fuel exergy, exergy losses, work outlet, irreversibilities, exergy destruction, exergy efficiency, entropy generation, and sustainable index) as well as the emissions characteristics of a diesel engine fueled with F-T diesel, ethanol, and biodiesel blends. To obtain a comprehensive comparison of the effects of combining F-T diesel and ethanol/biodiesel, the engine performance results were compared with the engine using diesel fuel, as baseline fuel. A steady engine condition was selected based on previous work of the research group [28]. Table 1. Summary of the energy and exergy results from other investigations. Ref. Year Engine Characteristics Operating Conditions Fuel Type Energy Efficiency (%) Exergy Efficiency (%) Outcomes [26] 2016 1-cylinder 4 stroke Antor 3LD510 12 different speeds (1000–3000 rpm) Full load Diesel D75-92/B10-20/E5 27.18–31.42 @ 1400 rpm 24.13–27.62 @ 2800 rpm 25.37–29.34 @ 1400 rpm 22.53–25.82 @ 2800 rpm ↑speed ↓efficiency ηdiesel >ηblends ψdiesel >ψblends [29] 2016 4-cylinders 4 stroke OM 314 Euro II 5 speeds 5 loads Diesel D50-78/B0-40/E0-16 n/a 25.21–32.88 @ 40% load 31.68–24.20 @ 80% load ↑speed ↓efficiency ↑load ↑efficiency [30] 2017 1-cylinder 4 stroke Kirloskar TV-1 1500 rpm 6 loads (20–120%) Diesel D30-45/B50/E5-20 23.49–27.05 @ 80% load 25.33–32.19 @ 100% load 20.91–24.49 @ 80% load 23.29–29.32 @ 100% load ↑load ↑efficiency ηD35/E15/B50 >ηdiesel ψD35/E15/B50 >ψdiesel [31] 2019 4-cylinders 4 stroke Yangdong Y85 1600 rpm Full load Diesel D41-78/B17-49/E510 25.82–27.08 n/a ηdiesel >ηblends [32] 2019 1-cylinder 4 stroke Kirloskar TAF1 1500 rpm 5 loads Diesel D90-98/B1.57.5/E0.5-2.5 19.25–21.24 @ 40% load 22.73–32.77 @ full load 21.45–28.21 @ full load ↑load ↑efficiency ηD90/B7.5/E2.5 >ηdiesel ψD90/B7.5/E2.5 >ψdiesel [33] 2020 6-cylinders 4 stroke Scania DC 1102 1450 rpm 5 loads (3–92%) GTL GTL25-75/B25-75 14.65–15.05 @ 3% load 43.93–45.27 @ 92% load 13.70–14.08 @ 3% load 41.12–41.65 @ 92% load ↑load ↑efficiency ηGTL25/B75 >ηGTL ψGTL25/B75 >ψGTL [34] 2020 1-cylinder 4 stroke NSB-8.18 1900 rpm 2 loads Diesel D82/B10/E8 21.9–22.0 @ 5.71 kW 22.6–22.9 @ 7.43 kW n/a ↑load ↑efficiency ηdiesel >ηD82/B10/E8 [27] 2020 1-cylinder 4 stroke 2 speeds Full load Diesel D60-75/B20/But5-20 30.17–31.92 @ 1400 rpm 26.91–27.83 @ 2800 rpm 28.13–29.77 @ 1400 rpm 25.03–25.96 @ 2800 rpm ↑speed ↓efficiency ηdiesel >ηblends ψdiesel >ψblends n/a: not available, D: diesel fuel, B: biodiesel fuel, E: ethanol fuel, GTL: GTL fuel, But: butanol fuel, FTD: Fischer-Tropsch diesel fuel. Appl. Sci. 2021,11, 5958 4 of 21 2. Materials and Methods 2.1. Experimental Setup The simplified schematic diagram of the experimental setup is shown in Figure 1. A common-rail single-cylinder direct injection diesel engine, four-stroke, water-cooled, with 84 mm ×90 mm of the cylinder bore and stroke respectively, connecting rod length of 160 mm, 16:1 of compression ratio, and a displacement of 499 cm 3 was used for this research. The fuel injection pressure could be set in the interval 500–1500 bar and the maximum indicated mean effective pressure (IMEP) could go up to 7 bar. An electric dynamometer was coupled to the engine, enabling the braking and the motoring of the engine. Appl. Sci. 2021, 11, x FOR PEER REVIEW 5 of 23 2. Materials and Methods 2.1. Experimental Setup The simplified schematic diagram of the experimental setup is shown in Figure 1. A common-rail single-cylinder direct injection diesel engine, four-stroke, water-cooled, with 84 mm × 90 mm of the cylinder bore and stroke respectively, connecting rod length of 160 mm, 16:1 of compression ratio, and a displacement of 499 cm3 was used for this research. The fuel injection pressure could be set in the interval 500–1500 bar and the maximum indicated mean effective pressure (IMEP) could go up to 7 bar. An electric dynamometer was coupled to the engine, enabling the braking and the motoring of the engine. Figure 1. Schematic diagram of the experimental setup. The fuels were tested with the engine kept at a steady-state condition of 2 bar IMEP, representing around 30% of the maximum load, and an engine speed of 1500 rpm. Before the test, the engine was warmed up to minimize the effects of emission variation during the engine cold-start. The IMEP was continuously monitored in order to keep a constant engine IMEP condition as the baseline fuel. All fuels were injected into the combustion chamber at 550 bar and two injection stages, divided into the pilot (15° before the top dead center, bTDC) and main (5° bTDC) injection. The pilot injection duration was kept constant (0.150 ms), but the main injection was adjusted during the change of fuels. Therefore, during shifting fuels, an adjustment of the main injection duration was necessary (0.499, 0.529, and 0.546 ms, respectively for Diesel, E15B35D50, and E15B35FTD50). Shell Global Solutions UK provided the ultra-low sulfur diesel fuel (<10 ppm sulfur) and the Fischer-Tropsch diesel. The ethanol was supplied by Fisher Scientific Company with a purity of 99.8%. The biodiesel was purchased from Egogas Ltd. (Kidderminster, UK) and the composition consisted of approximately 90% rapeseed methyl ester and 10% (v/v) palm oil methyl ester. The blends were prepared at the University of Birmingham and the physical and chemical properties of all fuels were calculated or obtained from the respective suppliers or publications [35,36], as shown in Table 2. Particularly, the diesel fuel that was used as a reference in this work was selected without any biodiesel in the composition (no oxygen content). The blends were prepared by mixing ethanol and biodiesel in the same volumetric proportion (15% of ethanol and 35% of biodiesel) with 50% F-T diesel (E15B35FTD50) and also with 50% diesel fuel (E15B35D50). Moreover, both fuel blends had the same oxygen content, and hence the influence of the oxygen when F-T diesel was blended with ethanol and biodiesel could be investigated. Figure 1. Schematic diagram of the experimental setup. The fuels were tested with the engine kept at a steady-state condition of 2 bar IMEP, representing around 30% of the maximum load, and an engine speed of 1500 rpm. Before the test, the engine was warmed up to minimize the effects of emission variation during the engine cold-start. The IMEP was continuously monitored in order to keep a constant engine IMEP condition as the baseline fuel. All fuels were injected into the combustion chamber at 550 bar and two injection stages, divided into the pilot (15 ◦ before the top dead center, bTDC) and main (5 ◦ bTDC) injection. The pilot injection duration was kept constant (0.150 ms), but the main injection was adjusted during the change of fuels. Therefore, during shifting fuels, an adjustment of the main injection duration was necessary (0.499, 0.529, and 0.546 ms, respectively for Diesel, E15B35D50, and E15B35FTD50). Shell Global Solutions UK provided the ultra-low sulfur diesel fuel (<10 ppm sulfur) and the Fischer-Tropsch diesel. The ethanol was supplied by Fisher Scientific Company with a purity of 99.8%. The biodiesel was purchased from Egogas Ltd. (Kidderminster, UK) and the composition consisted of approximately 90% rapeseed methyl ester and 10% (v/v) palm oil methyl ester. The blends were prepared at the University of Birmingham and the physical and chemical properties of all fuels were calculated or obtained from the respective suppliers or publications [ 35 , 36 ], as shown in Table 2. Particularly, the diesel fuel that was used as a reference in this work was selected without any biodiesel in the composition (no oxygen content). The blends were prepared by mixing ethanol and biodiesel in the same volumetric proportion (15% of ethanol and 35% of biodiesel) with 50% F-T diesel (E15B35FTD50) and also with 50% diesel fuel (E15B35D50). Moreover, both fuel blends had the same oxygen content, and hence the influence of the oxygen when F-T diesel was blended with ethanol and biodiesel could be investigated. Appl. Sci. 2021,11, 5958 5 of 21 Table 2. Physical and chemical properties of the fuels. Abbreviation % Volumetric Make-Up Diesel 100 diesel E15B35D50 15 ethanol + 35 biodiesel + 50 diesel E15B35FTD50 15 ethanol + 35 biodiesel + 50 F-T diesel Properties Diesel Ethanol Biodiesel F-T Diesel E15B35D50 E15B35FTD50 Chemical formula C14H26.1 C2H5OH C 19 H 35.3 O 2C16.89H35.77 C14.13H26.88O1.21 C15.52H31.53O1.24 Cetane number 53.9 8 54.7 79 47.7 (a) 59.6 a Heat of vaporization [kJ/kg] 243 858 216 339 [37] n/a n/a Lower heating value [MJ/kg] 43.11 26.83 37.8 43.90 38.86 a39.13 a Density at 15 ◦C [kg/m3]827.1 789.4 883.7 784.6 841.26 b820.01 b Aromatics [wt %] 24.4 0 ~0 0.3 n/a n/a Theoretical A-F ratio 14.56: 1 10.33: 1 11.77: 1 14.91: 1 12.92: 1 13.20: 1 C/H ratio 6.41 3.97 6.43 5.67 6.02 5.65 Carbon content [wt %] 86.47 52.14 77.15 84.91 78.06 77.28 Hydrogen content [wt %] 13.53 13.13 12.03 15.09 12.94 13.72 Oxygen content [wt %] 0 34.73 10.82 0 9.0 9.0 n/a: not available aEstimated based on the mass fraction for each component [38]. bEstimated based on the volumetric fraction for each component [38]. The O 2 emissions were measured using Testo 340 gas analyzer while the gaseous emissions of CO, CO 2 , NO x (NO and NO 2 ), nitrous oxide (N 2 O), ammonia (NH 3 ), formaldehyde (CH 2 O), total HC (sum of heavy and other unburned hydrocarbons), and individual light hydrocarbons species including methane (CH 4 ), ethane (C 2 H 6 ), acetylene (C 2 H 2 ), ethylene (C 2 H 4 ), propane (C 3 H 8 ), and propylene (C 3 H 6 ) were measured using an MKS MultiGas 2030 based on FTIR (Fourier Transform Infrared Spectroscopy) technique. For the calculation of the total hydrocarbons species, the corresponding flame ionization detector (FID) response factors were contained from the FTIR raw file and implemented for each individual HC reading as follows: THC = (1.1 × methane) + (2.4 × acetylene) + (1.9 × ethylene) + (2 × ethane) + (2.85 × propylene) + (1.35 × heavy HC) [ 39 ]. The exhaust gas sampling system of the emission analyzer was maintained at a constant temperature of 191 ◦ C by using a heated sampling line to prevent moisture and condensation during the sampling. The main technical characteristics and measurement accuracies are shown in Table 3. Table 3. Accuracy of the equipment used in this work. Measure Instrument Range Accuracy of the Measurement Range Exhaust gas (CO, CO2, NOx, THC) Multigas 2030 FTIR 10 ppb-100% full scale ±5% Exhaust gas (O2) Testo 340 0–25% ±0.2% Crank angle, engine speed Digital shaft encoder - ±1 rev/min In-cylinder pressure AVL GH13P pressure sensor 0–250 bar ±1% Temperature K-type thermocouples of 0–1250 ◦C±2.2 ◦C The engine was warmed up to minimize the effects of exhaust emission variation during the engine cold-start. Before the substitution between fuels, the fuel tanks and injection systems were cleaned. After this, the engine was kept constant for 30 min with the new fuel. The fuel consumption was measured in triplicate to obtain an average value. The FTIR measurements were recorded for 20 min to ensure reliability, while the Testo values were an average of two readings. Later, the experimental uncertainty has been calculated and error bars have been included in all the figures. Appl. Sci. 2021,11, 5958 6 of 21 2.2. Thermodynamic Analysis The control volume of the thermodynamic system is given in Figure 2, including the inlet and outlet terms. The equations used to calculate the inlet and outlet energy and exergy rates follow the literature [ 40 – 43 ]. The first and second laws of thermodynamics were applied to this system with the following assumptions: •The engine operation is studied at the steady-state condition; • The intake air and the outlet exhaust gases were considered as mixtures of ideal gases; • The environment (reference state) was considered at T 0 = 25 ◦ C and P 0 = 101.325 kPa and did not vary with time; • The kinetic and potential energy effects of incoming fluid streams and outgoing fluid streams were neglected [44]. Appl. Sci. 2021, 11, x FOR PEER REVIEW 7 of 23 values were an average of two readings. Later, the experimental uncertainty has been calculated and error bars have been included in all the figures. 2.2. Thermodynamic Analysis The control volume of the thermodynamic system is given in Figure 2, including the inlet and outlet terms. The equations used to calculate the inlet and outlet energy and exergy rates follow the literature [40–43]. The first and second laws of thermodynamics were applied to this system with the following assumptions:  The engine operation is studied at the steady-state condition;  The intake air and the outlet exhaust gases were considered as mixtures of ideal gases;  The environment (reference state) was considered at T0 = 25 °C and P0 = 101.325 kPa and did not vary with time;  The kinetic and potential energy effects of incoming fluid streams and outgoing fluid streams were neglected [44]. Figure 2. Control volume of the engine. 2.2.1. Energy Analysis Based on the assumptions made, the mass and energy balances of the control volume are given by Equations (1)–(3) below: ∑𝑚󰇗𝑖𝑛 =∑𝑚󰇗𝑜𝑢𝑡 (1) ∑E 󰇗nin =∑E 󰇗nout (2) E 󰇗nair +E 󰇗nfuel =E 󰇗nW+E 󰇗nexh +E 󰇗nloss (3) where 𝑚󰇗𝑖𝑛 and 𝑚󰇗𝑜𝑢𝑡 are respectively the inlet and outlet mass flow rates, ∑E 󰇗nin and ∑E 󰇗n𝑜𝑢𝑡 represents respectively the total energy inlet and outlet rates of the control volume, E 󰇗nair is the inlet energy rate of air, 𝐸𝑛 󰇗𝑓𝑢𝑒𝑙 is the inlet energy rate of the fuel, E 󰇗n𝑊 is the energy rate by work, 𝐸𝑛 󰇗𝑒𝑥ℎ is the energy outlet rate by the exhaust, and E 󰇗n𝑙𝑜𝑠𝑠 is the energy loss rate of the control volume. Because the intake air stream is at the same temperature as the reference state, the amount of energy inlet into the control volume can be neglected [45]. Figure 2. Control volume of the engine. 2.2.1. Energy Analysis Based on the assumptions made, the mass and energy balances of the control volume are given by Equations (1)–(3) below: ∑. min =∑. mout (1) ∑. Enin =∑. Enout (2) . Enair +. Enfuel =. EnW+. Enexh +. Enloss (3) where . min and . mout are respectively the inlet and outlet mass flow rates, ∑. Enin and ∑. Enout represents respectively the total energy inlet and outlet rates of the control volume, . Enair is the inlet energy rate of air, . Enf uel is the inlet energy rate of the fuel, . EnW is the energy rate by work, . Enexh is the energy outlet rate by the exhaust, and . Enloss is the energy loss rate of the control volume. Because the intake air stream is at the same temperature as the reference state, the amount of energy inlet into the control volume can be neglected [45]. The rate of energy inlet from the fuel ( . Enfuel ) to the control volume is calculated using the fuel mass flow ( . mfuel ) and the fuel lower heating value ( LHV ) as shown in Equation (4) below, where irepresents the different fuels that composed the blend. . Enfuel = 3 ∑ i=1 . mfueli·LHVfueli(4) Appl. Sci. 2021,11, 5958 7 of 21 The work rate ( . EnW ) was considered as the indicated power of the engine, which is expressed by Equation (5) below: . EnW=IMEP·Vd·N nR·60·103(5) where N(rpm) is the engine speed, Vd (m 3 ) represents the displaced volume, IMEP is the indicated mean effective pressure and nR is the number of crank revolutions for each power stroke per cylinder (e.g., nR is 2 for a four-stroke engine). The IMEP is calculated only from the cylinder pressure to represent the average pressure over a cycle of the engine. The outlet exhaust energy rate ( . Enexh ) of the control volume is determined as in Equation (6): . Enexh =∑ i . mi·hi(6) where . mi is the mass flow rate and hi is the enthalpy of each gaseous species in the engine exhaust, respectively. Then, the energy loss rate ( . Enloss ) of the control volume consists of all energy heat losses involved, which includes the heat transfers from cylinder walls (combustion chamber and piston), by coolant, and by the oil, except for the exhaust losses. In this work, the . Enloss is calculated as the difference between the energy inlet rate and the energy outlet rate (work and exhaust gases) from the control volume, such as in [ 26 , 46 ], as shown in Equation (7) below: . Enloss =. Enfuel −. EnW+. Enexh (7) Finally, the energy efficiency ( η ) of the control volume, based on the First Law of Thermodynamics (i.e., thermal efficiency), is defined as the work outlet ratio to the fuel energy inlet is shown in Equation (8). η= . EnW . Enfuel (8) 2.2.2. Exergy Analysis Similar to the case of the energy analysis, the same assumptions were valid for the exergy analysis of the control volume. The exergy balance can be expressed as Equation (9) as follows: ∑. Exin =∑. Exout +. Exdest (9) . Exair +. Exfuel =. ExW+. Exexh +. Exloss +. Exdest (10) where ∑. Exin , ∑. Exout and . Exdest represents respectively, the total exergy inlet and outlet rates and the exergy destruction (irreversibility) rate of the control volume. . Exair is the inlet exergy rate of air, . Exfuel is the inlet exergy rate from the fuel, . ExW is the exergy rate by work, . Exexh is the exergy outlet rate by the exhaust, and . Exloss is the exergy loss rate of the control volume. An equivalent assumption was considered for the air inlet rate, and by considering the environmental atmosphere as a reference state, the exergy associated with naturally aspirated air into the engine control volume is zero [47]. The inlet exergy rate from the fuel ( . Exfuel ) to the control volume can be determined as follows in Equation (11), where irepresents the different fuels that composed the blend. . Exfuel = 3 ∑ i=1 . mfueli·exfueli=. mfueli·LHVfueli·ϕfueli(11) where exfuel is the fuel-specific exergy of the fuel, which can be obtained by multiplying the fuel lower heating value by the chemical exergy factor ( ϕ ) of each fuel, which can be Appl. Sci. 2021,11, 5958 8 of 21 obtained through Equation (12) as in [ 41 , 42 ] for liquid fuels. The accuracy of this expression is estimated to be ±0.38%. ϕ=1.0401 +0.1728h c+0.0432o c+0.2169s c1−2.0628h c(12) where h,c,o, and sare respectively the mass fractions of hydrogen, oxygen, carbon, and sulfur of the fuel [41,42]. The exergy work rate ( . ExW ) is equal to the energy work rate of the control volume, as shown in Equation (13). . ExW=. EnW(13) The exergy rate of exhaust gases ( . Exexh) is composed of two components that are the physical (thermomechanical) and chemical exergies, which is expressed in Equation (14). . Exexh =∑ i . ni(exph,i+exch,i)(14) where . ni is the molar flow rate of each exhaust gas species, exph,i and exch,i are the specific physical and chemical exergies of each exhaust gas species, respectively. For a mass flow that goes through the volume control, the specific physical exergy rate (exph,i) of the exhaust gas species is obtained by Equation (15) below. exph,i=(hi−h0)−T0·(si−s0)(15) where siis the entropy of each gaseous species in the engine exhaust. For a gas mixture, the chemical exergy rate ( exch,i ) of the exhaust gas species can be calculated using Equation (16) as follow. exch,i=∑ i xi·εch,i+R·T0·∑ i (xi·lnγi·xi)(16) where xi is the molar fraction of each species of the exhaust gas, εch,i is the standard chemical exergy [ 41 , 42 ], R is the universal gas constant (8.314 kJ/kmol · K) and γi is the activity coefficient (γi= 1 for ideal gases). The exergy loss rate ( . Exloss ) shown in Equation (17) is considered to be the amount of exergy loss from the control volume, and the lost-exergy rate was assumed to be all heat losses occurring from the measured coolant temperature (T cool = 353.15 K) to the environment at reference state temperature, as in [29,48]. . Exloss =∑1−To Tcool . Enloss (17) Hence, the exergy destruction ( . Exdest ) is obtained from the exergy balance by Equation (18). . Exdest =. Exfuel −. ExW−. Exexh −. Exloss (18) Finally, the exergy efficiency ( ψ ) of the control volume is presented in Equation (19) as follows. ψ= . ExW . Exf uel (19) The entropy generation rate ( . Sgen ) is determined from the equation for the exergy destruction and is represented in Equation (20). . Sgen = . Exdest T0(20) Appl. Sci. 2021,11, 5958 9 of 21 The sustainability index (SI) is a function of exergy efficiency and is one of the methods to assess the sustainability of a system and measures its environmental impact and economics of energy technologies or resources, showing the effects of changing the exergy efficiency of a process. The SI has a key role in understanding the sustainable usage of fuel energy resources when dealing with the assessment of the system [ 25 , 27 ]. Thus, the SI can be calculated by Equation (21), as given in [ 49 ]. This relation shows how sustainability is affected by changing the exergy efficiency of a process. SI =1 1−ψ(21) 3. Results and Discussions 3.1. Gaseous Emissions The exhaust gaseous emissions results are shown in Figure 3. The overall lambda (i.e., actual air/fuel ratio over the stoichiometric air/fuel ratio) was found to be similar for the tested fuels as 2.9, 2.84, and 2.82 for diesel, E15B35D50, and E15B35FTD50, respectively. This is an indication that the difference between the fuels was the direct result of the fuel composition. The CO emissions with E15B35FTD50 and E15B35D50 increased compared to the reference fuel for the test condition. This could be a reflection of incomplete combustion occurring during the combustion process. In-cylinder pressure and detailed analysis of E15B35FTD50 and E15B35D50 fuel blends combustion process was reported in a previous work [ 50 ]. The lower combustion temperature due to the higher heat of vaporization of the ethanol is among the factors that might have influenced the higher CO results. Thus, the cooling effect of the alcohol dominates the effect of the total oxygen content of the blend of fuels, which might suppress the CO oxidation process [ 51 ]. Moreover, the lower cetane number of ethanol increases the time for ignition (longer ignition delay) of the air-fuel mixture [ 52 ]. Further, the higher viscosity of biodiesel has been indicated to promote poor atomization and hence increases CO emissions [ 53 ]. A recent study has reported that blends of F-T diesel and biodiesel resulted in slightly higher CO emission and the researchers attributed this to the insufficient evaporation and short mixing time of the blend during premixed combustion [ 13 ]. The application of a diesel oxidation catalyst could effectively reduce these CO levels from the combustion of these alternative fuels [50]. Appl. Sci. 2021, 11, x FOR PEER REVIEW 11 of 23 Figure 3. Exhaust gaseous emissions of diesel, E15B35D50, and E15B35FTD50. A slight decrease in CO2 emissions was observed when the engine was fueled with E15B35D50, while CO2 emissions for E15B35FTD50 are slightly higher than in the case of diesel fuel. This is in line with the carbon content of the estimated chemical formula of the fuel blends, though it has to be noted that the differences are within the confidence interval of the results. CO2 is a common product of the combustion of hydrocarbon fuels so the lower carbon content of ethanol molecules leads to lower CO2 formation, even on an energy basis (gCO2/MJfuel). In addition, for combustion stoichiometry, previous studies reported that an increase in CO2 emission with the addition of biofuels (ethanol and biodiesel) is attributed to the higher oxygen content of the oxygenated, which favors the combustion process and hence increases CO2 formation [54,55]. The combustion of E15B35D50 and E15B35FTD50 have reduced the THC emissions (mostly comprised of heavy unburned hydrocarbons) when compared to diesel fuel (Figure 3). It has been reported in the literature that ethanol and biodiesel addition to diesel blends increase [56] or decrease [14] the hydrocarbons emission. The total and the heavy hydrocarbons presented in the THC have decreased between 18%–33% for E15B35FTD50 in comparison with diesel or E15B35D50. This might be associated with the higher cetane number and the absence of aromatics of F-T diesel, contributing to the reduction in THC. Khan et al. [57] discussed that a higher cetane number and lower aromatic content are the main reasons for the occurrence of a shorter ignition delay time. Moreover, the virtually zero sulfur as well as the absence of aromatics, have been mentioned to result in a reduction in the engine-out exhaust emissions (i.e., reduce HC development and lowering PM) [58]. It has been previously reported in the literature that F-T diesel reduces hydrocarbons emissions [59] in comparison with diesel fuel. In this work, the addition of the oxygenated biofuels (ethanol and biodiesel) further reduced the unburned hydrocarbons emissions (Figure 3). The light-saturated HC (methane and ethane) and the unsaturated HC (acetylene, ethylene, and propylene) species, shown in Figure 4, have been separately analyzed from the heavy HC in order to provide an in-depth analysis of the THC. The heavy hydrocarbons solely correspond to approximately 88%, 85%, and 78% of THC for diesel, E15B35D50, and E15B35FTD50, respectively. The decrease in the heavy HC (seen in Figure 3) follows the same trend for the blends as with the THC. On the other hand, higher emissions of the light HC species (saturated and unsaturated) were measured for the E15B35FTD50 and E15B35D50 as compared to diesel fuel. This result could be attributed to the thermal decomposition of the alcohol component of the blends into shorter molecules of HC (i.e., light HC species) and CO, as discussed by Fayad et al.[28]. Nevertheless, the combustion of diesel is likely to produce heavier HC, which supports the reduction of the THC for the blends. 0 1 2 3 4 5 6 7 CO CO₂ / 100 O₂ / 100 Heavy HC THC NO NO₂ N₂O ×100 NOₓNH₃ ×100 Specific smissions (g/kWh) Diesel E15B35D50 E15B35FTD50 Figure 3. Exhaust gaseous emissions of diesel, E15B35D50, and E15B35FTD50. A slight decrease in CO 2 emissions was observed when the engine was fueled with E15B35D50, while CO 2 emissions for E15B35FTD50 are slightly higher than in the case of diesel fuel. This is in line with the carbon content of the estimated chemical formula of the fuel blends, though it has to be noted that the differences are within the confidence interval of the results. CO 2 is a common product of the combustion of hydrocarbon fuels so the Appl. Sci. 2021,11, 5958 16 of 21 Appl. Sci. 2021, 11, x FOR PEER REVIEW 18 of 23 (c) (d) (e) (f) Figure 8. Energy flow (Sankey) diagrams on the left side and exergy flow (Grossman) diagrams on the right side showing the comparison of energy and exergy efficiencies of the engine fueled with (a,b) diesel; (c,d) E15B35D50; (e,f) E15B35FTD50. Table 6 shows the summary of the energy and exergy efficiencies of the current work. The comparison of the present results with those by previous researchers, presented in Table 1, for different fuel types and engine operating conditions were either lower, nearly the same, or higher than that of the current work. The main reasons for the difference in these results are the fuel type, blend composition, and properties (e.g., cetane number, heating value, etc.) and the experimental conditions at which the diesel engines were evaluated. Table 6. Summary of the energy and exergy results of this work. Engine Characteristics Operating Conditions Fuel Type Energy Efficiency (%) Exergy Efficiency (%) 1-cylinder 4 stroke 1500 rpm 2 bar IMEP (30% load) Diesel E15/B35/D50 E15/B35/FTD50 26.2–26.6 24.4–24.9 ηdiesel > ηblends ψdiesel > ψblends D: diesel fuel, B: biodiesel fuel, E: ethanol fuel, FTD: Fischer-Tropsch diesel fuel. 4. Conclusions An investigation of the effects of alternative fuel blends (15% ethanol, 35% biodiesel, 50% diesel, E15B35D50, and 15% ethanol, 35% biodiesel, 50% F-T diesel, E15B35FTD50) on the energy and exergy analysis and exhaust emissions was performed. This work extended the understanding regarding the energetic and exergetic performance of synthesized diesel-like fuels and biofuels (ethanol and biodiesel) as well as the exhaust emissions from the combustion of alternative fuels. The energy and the exergy efficiencies were found to be similar, around 26% and 24%, respectively, with a little difference. The energy efficiency and also exergy efficiency of the diesel engine fueled with F-T diesel and ethanol/biodiesel (E15B35FTD50) was slightly lower than for the diesel fuel as a result of the inlet fuel energy and the higher fuel Figure 8. Energy flow (Sankey) diagrams on the left side and exergy flow (Grossman) diagrams on the right side showing the comparison of energy and exergy efficiencies of the engine fueled with ( a , b ) diesel; ( c , d ) E15B35D50; ( e , f ) E15B35FTD50. Table 6shows the summary of the energy and exergy efficiencies of the current work. The comparison of the present results with those by previous researchers, presented in Table 1 , for different fuel types and engine operating conditions were either lower, nearly the same, or higher than that of the current work. The main reasons for the difference in these results are the fuel type, blend composition, and properties (e.g., cetane number, heating value, etc.) and the experimental conditions at which the diesel engines were evaluated. Table 6. Summary of the energy and exergy results of this work. Engine Characteristics Operating Conditions Fuel Type Energy Efficiency (%) Exergy Efficiency (%) 1-cylinder 4 stroke 1500 rpm 2 bar IMEP (30% load) Diesel E15/B35/D50 E15/B35/FTD50 26.2–26.6 24.4–24.9 ηdiesel >ηblends ψdiesel >ψblends D: diesel fuel, B: biodiesel fuel, E: ethanol fuel, FTD: Fischer-Tropsch diesel fuel. 4. Conclusions An investigation of the effects of alternative fuel blends (15% ethanol, 35% biodiesel, 50% diesel, E15B35D50, and 15% ethanol, 35% biodiesel, 50% F-T diesel, E15B35FTD50) on the energy and exergy analysis and exhaust emissions was performed. This work extended the understanding regarding the energetic and exergetic performance of synthesized diesellike fuels and biofuels (ethanol and biodiesel) as well as the exhaust emissions from the combustion of alternative fuels. The energy and the exergy efficiencies were found to be similar, around 26% and 24%, respectively, with a little difference. The energy efficiency and also exergy efficiency of the diesel engine fueled with F-T diesel and ethanol/biodiesel (E15B35FTD50) was slightly lower than for the diesel fuel as a result of the inlet fuel energy and the higher fuel chemical exergy of the blend, respectively. The major cause of the inefficiency of the engine was the destroyed exergy caused by the irreversibilities of the system, especially by the combustion Appl. Sci. 2021,11, 5958 17 of 21 process. The exergy losses from exhaust gases and from the sum of the other losses also contributed to a decrease in the efficiency of the engine. Besides, it could be concluded that the sustainability index of the diesel fuel was higher than for the blends. The combustion of the fuel blends containing the biofuels (ethanol and biodiesel) produced lower levels of exhaust emissions such as HC, NO, NH 3 , and N 2 O compared to diesel fuel, however with a penalty on the CO emissions. Overall, the utilization of synthetic F-T diesel and biofuels such as ethanol and biodiesel can effectively mitigate the engine-out emissions, resulting in a similar efficiency of the engine. Thus, these alternative fuels that may be used without requiring modifications to the engine can also contribute to the reduction of fossil diesel utilization. These findings will be beneficial for biofuel and diesel hybrid electric vehicle synergies to achieve emission regulations by 2050. Besides, only limited literature evaluated the effects of the utilization of F-T diesel, ethanol, and biodiesel on the engine emissions without in-depth engine calibration. Furthermore, none included the light hydrocarbons speciation and unregulated emissions, such as NH 3 and formaldehyde. Due to the compatibility of advanced combustion strategies (i.e., RCCI and GCI) in hybrid electric vehicles, it is suggested that future research investigate the performance of ethanol and FT-diesel as well as higher chain alcohols (i.e., butanol and pentanol). Butanol and pentanol provide better miscibility as blended with FT-diesel. Additionally, investigation of advance combustion strategies of alternative fuels with aftertreatment systems can fasten hybridization. Author Contributions: Conceptualization, F.A.T., O.D., J.M.H. and A.T.; methodology, F.A.T., O.D. and J.M.H.; formal analysis, F.A.T., O.D. and R.L.; investigation, F.A.T., O.D., R.L. and R.P.; resources, A.T.; writing—original draft preparation, F.A.T.; writing—review and editing, O.D., J.M.H., A.T., J.M. and S.A.B.V.d.M.; supervision, A.T., J.M. and S.A.B.V.d.M.; project administration, S.A.B.V.d.M.; funding acquisition, S.A.B.V.d.M. All authors have read and agreed to the published version of the manuscript. Funding: This research received no external funding. Institutional Review Board Statement: Not applicable. Informed Consent Statement: Not applicable. Data Availability Statement: Not applicable. Acknowledgments: F.A.T. gratefully acknowledges the managers and researchers of the Future Engines & Fuels Laboratory of the University of Birmingham for the kind assistance, technical support, lab facilities, and reviews during the hosted research period of time. The Coordenação de Aperfeiçoamento de Pessoal de Nível Superior-Brasil (CAPES)-Finance Code 001 is also acknowledged for the scholarship (No. 88887.199341/2018-00) to pursue F.A.T. postgraduate studies under the International Cooperation Program with the University of Birmingham and CAPES (Ref. 40/2014). Conflicts of Interest: The authors declare no conflict of interest. Abbreviations BTL Biomass-to-liquid bTDC Before the top dead center C2H2Acetylene C2H4Ethylene C2H6Ethane C3H6Propylene C3H8Propane CH2O Formaldehyde CH4Methane CI Compression ignition CO Carbon monoxide CO2Carbon dioxide Appl. Sci. 2021,11, 5958 18 of 21 E15B35D50 15% ethanol, 35% biodiesel, 50% diesel E15B35FTD50 15% ethanol, 35% biodiesel, 50% Fischer-Tropsch diesel FID Flame ionization detector F-T Fischer-Tropsch FTD Fischer-Tropsch diesel FTIR Fourier transform infrared GTL Gas-to-liquid HC Hydrocarbons HEV Hybrid electric vehicles ICE Internal combustion engines IMEP Indicated mean effective pressure LHV Lower heating value N2O Nitrous oxide NH3Ammonia NO Nitrogen oxige NO2Nitrogen dioxide NOxNitrogen oxides PCIS Pre-chamber ignition system PM Particulate matter RCCI Reactivity-controlled compression ignition SI Sustainability index TDC Top dead center THC Total hydrocarbons Symbols . En energy rate, W . Ex exergy rate, W ex specific exergy, J/kg or J/mol henthalpy, J/kg or J/mol . mmass flow rate, kg/s . nmolar flow rate, mol/s Nengine speed, rpm nRnumber of crank revolutions for each power stroke per cylinder, - P absolute pressure, Pa T temperature, ◦C or K Vddisplaced volume, m3 Runiversal gas constant, J/kg·K or J/mol·K sentropy, J/kg·K or J/mol·K . SEntropy rate, W xmolar fraction, - Subscripts 0reference state ch chemical cool coolant dest destruction exh exhaust gen generation iindividual gaseous species in inlet out outlet ph physical Wwork Greek sumbols γactivity coefficiency, - εstandard chemical exergy, J/mol ηenergy efficiency, - ϕchemical exergy factor, - ψexergy efficiency, - Appl. Sci. 2021,11, 5958 19 of 21 References 1. Correa, G.; Muñoz, P.M.; Rodriguez, C.R. A Comparative Energy and Environmental Analysis of a Diesel, Hybrid, Hydrogen and Electric Urban Bus. Energy 2019,187, 115906. [CrossRef] 2. Serrano, J.R.; Novella, R.; Piqueras, P. 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