Accepted Manuscript Investigating the effect of using diethyl ether as a fuel additive on diesel engine performance and combustion Amr Ibrahim PII: DOI: Reference:
S1359-4311(16)31192-9 http://dx.doi.org/10.1016/j.applthermaleng.2016.07.061 ATE 8659
To appear in:
Applied Thermal Engineering
Received Date: Revised Date: Accepted Date:
23 May 2016 27 June 2016 9 July 2016
Please cite this article as: A. Ibrahim, Investigating the effect of using diethyl ether as a fuel additive on diesel engine performance and combustion, Applied Thermal Engineering (2016), doi: http://dx.doi.org/10.1016/ j.applthermaleng.2016.07.061
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Investigating the effect of using diethyl ether as a fuel additive on diesel engine performance and combustion Amr Ibrahim* Mechanical Engineering Department, Alexandria University, Alex 21544, Egypt
Abstract The diethyl ether (DEE) is a renewable oxygenated fuel, which has favorable characteristics to be used as a fuel additive for the diesel engines. The aim of this study was to experimentally investigate the effect of blending the DEE with the diesel fuel in different proportions up to 15% by volume on diesel engine performance, combustion characteristics, and engine stability. All the tests were conducted using a single-cylinder direct-injection diesel engine without modification at a fixed engine speed of 1500 rpm and variable load conditions. It was found that using the DEE as a fuel additive improved the engine performance significantly for the most of engine load conditions. The engine maximum brake thermal efficiency increased by 7.2% and the lowest brake specific fuel consumption decreased by 6.7% when 15% of DEE was used in the fuel blend compared to the diesel fuel. In addition, using the DEE increased the maximum cylinder pressure and maximum net heat release rate compared to the diesel fuel for the most of engine load conditions. Engine stability and combustion duration slightly reduced while the start of *
Corresponding Author: Tel: +2 01091320421, Email:
[email protected] , Address: Mechanical Engineering Department, Alexandria University, Alex 21544, Egypt.
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combustion timing changed insignificantly when the DEE fuel blends were used as alternatives to the diesel fuel. Keywords: combustion; diesel; diethyl ether; DEE; engine
1. Introduction The depletion of fossil fuels and the increase of emission restrictions have led to exploration of cleaner renewable fuels that can be used for the diesel engines [1]. Most of these fuels are used as fuel additives to the conventional diesel fuel with different proportions in order to reduce engine emissions without a significant deterioration in engine performance [2, 3]. These fuels include the oxygenated fuels such as alcohols. Although alcohols are used as alternative fuels for the spark-ignition engines due to their high octane number, they are currently being tested to be used as fuel additives for the diesel engines as well [4]. Ethanol and butanol are the dominant alcohol types that are extensively tested as potential oxygenated fuel additives for the diesel engines. However, ethanol and butanol have significantly lower cetane number than the diesel fuel [4], which can imply that ethanol and butanol may not be the optimum alcohol types that can be used as fuel additives for the diesel engines. Alternatively, the diethyl ether (DEE) has a higher cetane number compared to the diesel fuel and higher calorific value compared to both ethanol and butanol [5-8]. The DEE is a biofuel as it can be produced from ethanol via a dehydration process using acid catalysts [9-10]. Similarly, the dimethy ether (DME) can be produced from the dehydration of methanol. However, the DEE is a liquid
2
fuel at the ambient temperature, which makes it more suitable than the DME, which is a gaseous fuel, to be used as a fuel additive for diesel engines without modifications [6-8]. The chemical formula of DEE is C4H10O, which makes the oxygen content in the fuel equal to 21.6% by mass [11]. Table 1 compares the properties of DEE with the corresponding properties of conventional diesel fuel. Table 1 indicates that the DEE has favorable fuel characteristics for the diesel engines. These characteristics include its high cetane number, low auto-ignition temperature, high volatility, and high oxygen content. In addition, the DEE has a high miscibility with the diesel fuel and broad flammability limits [11-12]. Although the DEE was previously used as a sole fuel for cold start aid of the diesel engines due to its low auto-ignition temperature [13], the extremely low viscosity of DEE can increase the wear rate in the diesel injection system, which restricts using the DEE as a sole sustainable fuel for the diesel engines without modifications. Therefore, it is favorable to use the DEE only as a fuel additive by blending it in limited proportions with the diesel fuel. Also, there are some concerns on the impact of DEE on air pollution during storage due its high volatility and its tendency to oxidize forming peroxides in storage [13]. There are a limited number of studies that investigated using the DEE as a supplement fuel in the diesel engines. Rakopols and coworkers [6-8] demonstrated that blending the DEE with the diesel fuel up to a maximum percentage of 24% decreased engine smoke, CO, and NOx emissions and increased engine HC emissions. It was also found that the engine fuel
3
consumption slightly increased and the engine brake thermal efficiency was unchanged. In addition, it was shown that using the DEE as a fuel additive did not affect engine stability and resulted in acceptable cycle to cycle variations [8]. Banapurmath and coworkers [14] concluded that the addition of DEE to the diesel fuel up to 20% increased engine brake thermal efficiency and NOx emissions and decreased engine smoke, CO, and HC emissions. Paul and coworkers [15] studied the effect of blending the DEE with the diesel fuel in two different proportions on diesel engine performance and emissions. It was found that using 5% of DEE in the fuel blend increased engine brake thermal efficiency and NOx emissions and decreased engine brake specific energy consumption. On the other hand, increasing the percentage of DEE to 10% decreased the engine thermal efficiency and NOx emissions and increased engine brake specific energy consumption. However, using both 5% and 10% of DEE led to a decrease in engine smoke and HC emissions compared to the diesel fuel. Murat and coworkers [11] investigated using the DEE as a fuel additive in a diesel engine operating in a dual fuel mode containing 40% natural gas. The DEE was blended with the diesel fuel with two proportions of 5% and 10% and used as a pilot fuel. The authors concluded that using the DEE as a fuel additive resulted in an improvement in engine brake thermal efficiency and specific energy consumption and reduced engine CO and NO emissions compared to the use of standard dual fuel mode. In addition, it was demonstrated that using the DEE as a fuel additive to diesel-biodiesel
4
mixtures can result in a decrease engine fuel consumption, and engine smoke and CO emissions [12, 16]. It can be concluded from the previous studies that the DEE can be used as a fuel additive in order to improve diesel engine performance and reduce emissions. However, there are only a limited number of investigations that studied using the DEE as a fuel additive for diesel engines. Therefore, more investigations are needed in order to fully assess the effect of using the DEE as a renewable supplement fuel on diesel engine performance, combustion characteristics, engine stability, and emissions using different engine specifications and operating conditions. Similar conclusion was also made by other studies [6-8]. Therefore, the aim of this study was to experimentally investigate the effect of using the DEE as a fuel additive on diesel engine performance, combustion characteristics, and engine stability. The DEE was blended with the diesel fuel with different proportions up to 15% and tested in a single cylinder diesel engine without modification. Most of the studies which previously investigated using the DEE in diesel engines were conducted at engine speeds either lower or higher than 1500 rpm [5-8, 12, 16]. However, diesel engines which are used for electrical power generation usually operate with a fixed speed of 1500 rpm. Therefore, in the current study, all the tests were conducted at a constant engine speed of 1500 rpm and different engine load conditions varying from low to full load condition.
5
2. Methodology and Experimental setup The experimental research was carried out using a single-cylinder, fourstroke, direct- injection, air-cooled, TecQuipment TD212 diesel engine. The engine was naturally aspirated and had a cylinder head modified for the installation
of
an
in-cylinder
pressure
transducer.
The
engine
specifications are shown in Table 2. The engine was coupled to the TecQuipment TD200 engine test set. The engine test set included a hydraulic dynamometer to control the engine load as shown schematically in Figure 1. The fuel flow rate was measured using the DVF1 automatic volumetric fuel gauge with a digital readout of flow rate, which was a pipette fitted with optical sensors. The inlet air flow rate was measured using an orifice plate. A k-type thermocouple and a differential pressure transducer were fitted near the orifice plate in order to measure the air temperature and pressure; respectively, needed for the inlet air flow rate calculation. The orifice plate was installed near an air box in order to dampen the inlet air flow pulses. The engine torque was measured using a load cell while the engine speed was measured using an optical sensor. All the test measured variables such as the engine torque, speed, power, flow rates, temperatures, etc were digitally displayed on instrument modules. In addition, the TecQuipment Versatile Data Acquisition System (VDAS) was used in order to accurately monitor and record all the measured data on a computer. The in-cylinder pressure was measured using a piezoelectric pressure transducer (ECA 101) and recorded against the crank angle, which was measured using a shaft encoder (ECA 102). The signals of both the pressure
6
transducer and shaft encoder were received by an engine cycle analyzer (ECA 100) as shown in Figure 1. The engine cycle analyzer was a hardware unit with a charge amplifier and signal conditioning circuits.
The engine
cycle analyzer was connected to a computer via dedicated software in order to display and record the pressure-crank angle data. Three different proportions of DEE were mixed with the diesel fuel as shown in Table 3. These proportions were 5%, 10%, and 15% (by volume). Samples of the all tested fuels were sent to the chemical laboratory of Alexandria Petroleum Company in order to measure the fuel density and gross calorific value using the ASTM D-4052 and ASTM D-4868 standard methods; respectively. These properties are shown in Table 4.
The net heat release rate,
dQ , was calculated as a function of the measured d
in-cylinder pressure, p, and the cylinder volume, V, using the following equation [17]:
dQ γ dV 1 dp p V dθ γ 1 dθ γ 1 dθ
(1)
Where is the crankangle and is the specific heat ratio of the in-cylinder contents. Equation 1 was derived by applying the energy conservation principle on the in-cylinder contents during the combustion process using a single zone combustion model [17].
3. Results and discussion In this section, the effect of mixing the DEE in different proportions up to 15% (by volume) with the diesel fuel on diesel performance, combustion
7
characteristics, and engine stability is discussed. All the tests were conducted at a constant engine speed of 1500 rpm and variable load conditions, which varied from low to full load condition. Engine brake torque varied from the lowest load of 1 Nm to the full load of 9 Nm with corresponding brake and indicated power ranges of 0.157 to 1.413 kW and 0.7 to 2 kW; respectively. 3.1 Thermal efficiency and bsfc Figures 2 and 3 show the variations of engine brake thermal efficiency and brake specific fuel consumption (bsfc) with engine torque; respectively, for different fuels. Both Figures 2 and 3 indicate that using the DEE as a fuel additive to the diesel fuel improved the engine performance significantly as the engine brake thermal efficiency increased and the engine bsfc decreased for the most of engine load conditions. The engine maximum thermal efficiency increased from 32% for the diesel fuel to 32.5%, 34.3%, and 34.3% when the DEE was blended with the diesel fuel with proportions of 5%, 10%, and 15%; respectively. On the other hand, the lowest engine bsfc decreased from 0.252 kg/kWh to 0.247, 0.235, and 0.235 kg/kWh when the DEE was blended with the diesel fuel with proportions of 5%, 10%, and 15%; respectively. That means using the DEE as a fuel additive with a proportion of 15% increased the engine maximum brake thermal efficiency by 7.2% and decreased the lowest engine bsfc by 6.7% compared to the diesel fuel. The presence of oxygen in the DEE helped in improving the combustion efficiency and burning the fuel more completely. Also, the high volatility of DEE improved the fuel-air mixing prior to combustion, which further increased the combustion efficiency. In addition, the air to fuel ratio
8
slightly increased when the DEE was blended with the diesel fuel in different proportions for the most of engine load conditions as indicated in Figure 4 due to the lower density of DEE compared to the diesel fuel. Furthermore, the stoichiometric air to fuel ratio of the DEE is lower compared to the diesel fuel as shown in Table 1. The higher operating air to fuel ratio and the lower stoichiometric air to fuel ratio of the DEE resulted in the engine to operate with a leaner fuel-air mixture when the DEE was blended with the diesel fuel in different proportions, which further improved the combustion efficiency causing an increase in engine thermal efficiency and a decrease in engine bsfc. The improvement in both engine thermal efficiency and bsfc was more significant when the percentage of diethyl ether increased in the fuel blend during engine higher load operation compared to lower load conditions. That was because the engine operated with a richer fuel-air mixture during engine higher load operation as indicated in Figure 4. Hence, the presence of high volatility oxygenated fuel improved the combustion efficiency more significantly during engine higher load operation. It is worth mentioning that when the percentage of DEE increased to 20% in the fuel blend, engine instability occurred as the engine was difficult to stabilize at a fixed engine speed and load condition especially at the higher engine load conditions. Such phenomenon was also observed by other studies [7]. This engine instability was possibly due to the high volatility of DEE which created vapor bubbles and caused vapor lock in the fuel lines when the DEE was used with a relatively higher proportion in the fuel blend [7].
9
As using the two blends of D90 and D85 improved the engine performance more significantly compared to the diesel and D95 fuels, the effects of using only the D90 and D85 fuels on engine combustion characteristics and engine stability compared to the diesel fuel will be discussed in the next sections. It was found that removing the D95 fuel results from the relevant figures could help the reader to understand more clearly and adequately the effect of using the DEE as fuel additive on combustion characteristics and engine stability without changing the general trend of the results. 3.2 Specific heat ratio The heat release rate was calculated using Equation (1). This equation was derived by applying the first law of thermodynamics on the engine cylinder during the combustion process using a single zone combustion model [17]. γ represents the specific heat ratio of the in-cylinder contents during the combustion process. The net heat release rate was calculated in most of the studies found in the literature by assuming a constant value of γ of usually in the range from 1.3 to 1.35 [4, 17-19]. However, γ varies with the in-cylinder temperature and composition [4]. Heywood [17] demonstrated that the value of γ at the start of combustion can be as high as 1.35, which is comparable to the hot air specific heat ratio at the end of the compression process while γ can be as low as 1.26 by the end of combustion process, which is comparable to the specific heat ratio of the burned gas formed by the end of combustion. Using inappropriate values for γ can result in significant errors in the engine heat release rate calculations [19]. Also, assuming a constant value of γ for the all the engine operation conditions and tested fuels can lead to inaccurate
10
results when the engine combustion characteristics are analyzed [18-19]. Although using the instantaneous values of γ calculated as a function of the engine crankangle during the combustion process can reduce the errors associated with the heat release rate calculations, it was demonstrated that selecting an appropriate average value for γ during the combustion process can lead to satisfactory results [19]. However, this average value of γ must vary with the change of engine operating conditions. Abbaszadehmosayebi and coworkers [18-19] described a methodology to calculate the average specific heat ratio during the combustion process from the measured in-cylinder pressure data. As the compression of the in-cylinder contents before the combustion is close to the isentropic process, the isentropic relationship of
=constant can be applied through the process,
where γu represents the specific heat ratio of the in-cylinder contents during the compression process [17]. Similarly, the isentropic relationship of =constant can be applied through the expansion process after the end of combustion, where γb represents the specific heat ratio of the in-cylinder contents during the expansion process. Therefore, plotting the log p-log V relationships from the start of compression process to the start of combustion, and from the end of combustion to the end of expansion process results in two approximate straight lines with the slopes of γu and γb; respectively [17]. Abbaszadehmosayebi [18-19] showed that the average of both γu and γb could represent an appropriate value for γ, which was used for the heat release rate calculation as indicated in Equation (1). Such methodology was used in the current paper in order to calculate γ from the measured in-cylinder pressure data at different engine load conditions for all the tested fuels.
11
Figure 5 shows the variation of specific heat ratio with engine indicated power for the tested fuels. The specific heat ratio decreased with the increase of engine indicated power because the engine power was increased by increasing the amount of burned fuel within the cylinder, which increased the in-cylinder pressure, as shown in Figure 6, and temperature resulting in a decrease in the in-cylinder specific heat ratio. A similar trend was also demonstrated by Heywood [17], who illustrated that the specific heat ratios of both the unburned and burned in-cylinder gases decreased with the increase of in-cylinder temperature. Also, Figure 5 shows that blending the DEE with the diesel fuel resulted in a decrease in the specific heat ratio of the cylinder contents during combustion for the most of engine load conditions compared to the diesel fuel. Blending the DEE with diesel fuel increased the maximum heat release rate as discussed in the next section. The increase of the maximum heat release rate was expected to be associated with an increase in the cylinder temperature, which resulted in a decrease in the specific heat ratio. 3.3 Combustion characteristics Figures from 7 to 9 show the rate of change of cylinder pressure versus crankangle for different fuels at the engine low, half, and full load conditions; respectively. The relevant net heat release rate variations induced for different fuels during engine low, half, and full load conditions are shown in Figures from 10 to 12; respectively. The engine low, half, and full loads were relevant to engine brake torque of 1, 5, and 9 Nm; respectively, with a corresponding engine indicated power of 0.7, 1.4, and 2 kW; respectively.
12
The combustion started when both the rate of pressure rise and net heat release rate started to rapidly increase after the start of fuel injection, which occurred at about 350 degrees (10 degrees before the TDC). The combustion started at almost 355.2, 354.2, and 354.2 degrees at the engine low, half, and full load conditions; respectively, for the all tested fuels. Although the DEE has much higher cetane number than the diesel fuel, the cetane number of the diesel-DEE blends could be comparable to the cetane number of the diesel fuel. It was reported that the DEE ignition was inhibited by the diesel fuel and that adding the DEE to the diesel fuel may decrease the cetane number of the blend to be lower than the cetane number of the diesel fuel as the DEE could interact with the aromatics in the diesel fuel delaying the onset of ignition [13]. However, Paul and coworkers [15] demonstrated that the cetane number of diesel fuel increased from 49 to 53.75 and 57.5 when the diesel fuel was blended with the DEE with proportions of 5%, and 10%; respectively. Figures 10 to 12 indicate that the change of fuel type did not significantly affect the start of combustion timing. However, the start of combustion timing was advanced by almost 1 degree when the engine load increased from low to full load condition. The engine load was increased by increasing the amount of burned fuel inside the cylinder, which increased the maximum engine cylinder pressure (as shown in Figure 6) and temperature causing a decrease in the ignition delay period and earlier start of combustion. The amount of fuel accumulated during the ignition delay period and mixed with air was rapidly burned immediately after the start of combustion during the premixed combustion phase causing a rapid increase in the rate of in-
13
cylinder pressure rise and net heat release rate. The maximum heat release rate induced during the premixed combustion phase occurred at almost the same crankangle at which the maximum rate of pressure rise occurred for each engine load condition. Figures 7 to 12 indicate that the maximum rate of heat that released during the premixed combustion phase and the associated rate of cylinder pressure rise decreased with increasing engine load. Increasing the engine load decreased the ignition delay period; and consequently, the amount of accumulated fuel during the delay period decreased causing a decrease in the maximum rate of heat released during the premixed combustion phase. Figures 7 to 12 indicate that blending the DEE with diesel fuel caused an increase in the maximum rate of cylinder pressure rise and maximum rate of heat released during the premixed combustion phase. The higher volatility of the DEE, which improved the fuel-air mixing, and the presence of oxygen in the fuel improved the combustion efficiency, which caused an increase in the heat released during the premixed combustion phase. The mixing controlled combustion phase started immediately after the end of the premixed combustion phase. The end of the premixed combustion phase occurred at 361.5, 360.5, and 359.5 degrees at the engine low, half, and full load conditions; respectively. The amount of fuel burned in the cylinder during the mixing controlled combustion phase increased with increasing the engine load. Therefore, the amount of heat released during the mixing controlled combustion phase and the associated rate of cylinder pressure rise increased and became more significant with increasing the engine load as
14
indicated in Figures from 7 to 12. The higher volatility of the DEE increased the fuel-air mixing rate during the mixing controlled combustion phase which caused an increase in the maximum heat release rate induced during the mixing controlled phase for the DEE fuel blends compared to the diesel fuel. The end of combustion was identified when the net heat release rate became almost zero by the end of the mixing controlled combustion phase. Figure 10 shows that the combustion ended at the engine low load condition for the fuels D100, D90, and D85 at 390.2, 389.8, and 389 degrees; respectively. The end of combustion occurred at the half load condition for the fuels D100, D90, and D85 at 396, 395, and 395 degrees; respectively as shown in Figure 11. On the other hand, Figure 12 indicates that the combustion ended at the engine full load condition for the fuels D100, D90, and D85 at 403, 402, and 400 degrees; respectively. It can be concluded that the combustion duration increased significantly with increasing engine load. The combustion duration increased by almost 10 crankangle degrees when the engine load increased from low to full load condition for all fuels. This was because the amount of fuel burned within the cylinder increased with increasing engine load; and consequently, more crankangle duration was needed for the extra amount of fuel to be burned. On the other hand, using the DEE fuel blends as alternatives to the diesel fuel increased the maximum heat release rate; and consequently, the combustion duration slightly decreased. 3.4 Coefficient of variation (COV) The engine stability was assessed for the tested fuels by calculating the coefficient of variation in indicated mean effective pressure (COV). Cycle-to-
15
cycle variations are observed when the in-cylinder pressure is measured for more than one thermodynamic cycle. Cycle-to-cycle pressure variation is mainly caused by the cycle-to-cycle variation in the combustion process [17]. One important measure of the cyclic variability, which can be calculated from the measured in-cylinder pressure data, is the coefficient of variation in indicated mean effective pressure (COV ) which can be calculated as follows [20-21]:
Where
is the average indicated mean effective pressure calculated for
a number of cycles, n, while
is the standard deviation in indicated
mean effective pressure. Both parameters can be calculated as follows [2021]:
Heywood [17] demonstrated that the engine stability was usually affected when the COV, exceeded about 10%. However, other studies showed that the engine stability started to deteriorate when the COV increased above 5% [21]. Figure 13 shows the variations of COV with engine indicated power for different fuels. The COV was calculated for five thermodynamic cycles. Figure 13 shows that the COV is generally higher for the engine lower load conditions compared to higher load conditions especially for the D85 fuel. That was because the engine was operated with a leaner fuel-air mixture
16
during engine lower load operation as indicated in Figure 4. Operating the engine with leaner fuel-air mixtures can increase the engine cycle to cycle variations [22]. The COV was below 5% for the most of engine load conditions especially at higher engine loads for the all tested fuels. However, the average COV increased slightly when the DEE fuel blends were used as alternatives to the diesel fuel. The average COV for D100, D90, and D85 was 2.85%, 2.98%, and 3.391%; respectively. Increasing the percentage of DEE in the fuel blend increased the chance of creating fuel vapor bubbles in the fuel tubes due to the high volatility of diethyl ether causing vapor lock, which decreased engine stability, and consequently; the COV increased.
4. Conclusions In this study, the performance, combustion characteristics, and stability of a diesel engine fuelled by different blends of DEE and diesel fuels, with a maximum DEE proportion of 15% by volume, were experimentally investigated and compared at a constant engine speed of 1500 rpm and different engine load conditions. The results indicated that the DEE fuel is strongly recommended to be used as an oxygenated renewable fuel additive for the diesel engines without modifications as its use improved the engine performance significantly and resulted in a tolerable change in engine stability and combustion characteristics compared to the diesel fuel for the most of engine load conditions. The maximum engine brake thermal efficiency increased by 7.2% while the lowest brake specific fuel consumption decreased by 6.7% when the DEE was used a fuel additive with
17
a proportion of 15% compared to the diesel fuel. It was also found that using the DEE fuel increased the maximum cylinder pressure and maximum heat release rate compared to the diesel fuel for the most of engine load conditions. In addition, the combustion duration slightly reduced and the start of combustion timing changed insignificantly. Although the average COV increased slightly with increasing the proportion of DEE in the fuel blend up to 15%, the overall engine stability was satisfactory. However, when the proportion of DEE increased to 20%, the engine stability deteriorated as the engine was difficult to stabilize at a fixed speed and load condition.
References [1] Devaraj J, Robinson Y, Ganapathi P. Experimental investigation of performance, emission and combustion characteristics of waste plastic pyrolysis oil blended with diethyl ether used as fuel for diesel engine. Energy 2015; 85: 304-9 [2]El-Adawy M, Ibrahim A, El-Kassaby MM. An experimental evaluation of using waste cooking oil biodiesel in a diesel engine. Energy Technol 2013; 1(12):726-34. [3]Ibrahim A, El-Adawy M, El-Kassaby MM. The impact of changing the compression ratio on the performance of an engine fuelled by biodiesel blends. Energy Technol 2013; 1(7):395-404. [4]Ibrahim A. Performance and combustion characteristics of a diesel engine fuelled by butanol-biodiesel-diesel blends. Appl Therm Eng 2016; 103:65159
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[5]Rakopoulos DC, Rakopoulos CD, Giakoumis EG. Impact of properties of vegetable oil, bio-diesel, ethanol and n-butanol on the combustion and emissions of turbocharged HDDI diesel engine operating under steady and transient conditions. Fuel 2015; 156:1–19. [6]Rakopoulos DC, Rakopoulos CD, Giakoumis EG, Papagiannakis RG, Kyritsis DC. Influence of properties of various common bio-fuels on the combustion and emission characteristics of high-speed DI (direct injection) diesel engine: Vegetable oil, bio-diesel, ethanol, n-butanol, diethyl ether. Energy 2014; 73: 354-66. [7]Rakopoulos DC, Rakopoulos CD, Giakoumis EG, Dimaratos AM. Characteristics of performance and emissions in high-speed direct injection diesel engine fueled with diethyl ether/diesel fuel blends. Energy 2012; 43:214-24 [8]Rakopoulos DC, Rakopoulos CD, Giakoumis EG, Dimaratos AM. Studying combustion and cyclic irregularity of diethyl ether as supplement fuel in diesel engine. Fuel 2013; 109:325–35 [9]Barik D, Murugan S. Effects of diethyl ether (DEE) injection on combustion performance and emission characteristics of Karanja methyl ester (KME)–biogas fueled dual fuel diesel engine. Fuel 2016; 164: 286–96 [10]Sezer I. Thermodynamic, performance and emission investigation of a dieselengine running on dimethyl ether and diethyl ether. Int J Therm Sci 2011; 50:1594-1603
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[11]Karabektas M, Ergen G, Hosoz M. The effects of using diethylether as additive on the performance and emissions of a diesel engine fuelled with CNG. Fuel 2014; 115:855–60 [12]Imtenan S, Masjuki HH, Varman M, Fattah IMR, Sajjad H, Arbab MI. Effect of n butanol and diethyl ether as oxygenated additives on combustion– emission-performance characteristics of a multiple cylinder diesel engine fuelled with diesel–jatropha biodiesel blend. Energy Convers Manage 2015; 94: 84–94 [13]Balley B, Eberhardt J, Goguen S, Erwin J.Diethyl ether (DEE) as a renewable diesel fuel. SAE paper 972978 [14]Banapurmath NR , Khandal SV, Swamy RL,
Chandrashekar TK.
Alcohol (Ethanol and Diethyl Ethyl Ether)-Diesel Blended Fuels for Diesel Engine Applications-A Feasible Solution. Adv Automob Eng 2015; 4:117 [15]Paul A, Bose PK, Panua RS, Debroy D. Study of performance and emission characteristics of a single cylinder CI engine using diethyl ether and ethanol blends.J Energy Inst 2015; 88: 1-10 [16]Qi DH, Chen H, Geng LM, Bian YZ. Effect of diethyl ether and ethanol additives on the combustion and emission characteristics of biodiesel-diesel blended fuel engine. Renew Energy 2011; 36: 1252-58 [17]Heywood JB. Internal combustion engine fundamentals. International ed. New York: McGraw-Hill Book Company; ISBN 0071004998, 1988
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[18]Abbaszadehmosayebi G. Diesel engine heat release analysis by using newly defined dimensionless parameters. PhD Thesis 2014; School of Engineering and Design, Brunel University, United Kingdom.
[19]Abbaszadehmosayebi G, Ganippa L. Determination of specific heat ratio and error analysis for engine heat release calculations. Appl Energy 2014; 122:143-50. [20] Ibrahim A. Bari S. Effect of varying compression ratio on a natural gas SI engine performance in the presence of EGR. Energy Fuels 2009; 23:494956. [21]Ibrahim A, Bari S. An experimental investigation on the use of EGR in a supercharged natural gas SI engine. Fuel 2010; 89:1721-30. [22]Stone R. Introduction to internal combustion engines. 4 th ed. Palgrave Macmillan; ISBN 978-0-230-57663-6, 2012 List of table captions Table 1. Properties of diesel vs. DEE Table 2. Engine specifications Table 3. Types of tested fuels Table 4. Properties of tested fuels List of figure captions Figure 1. A schematic of experimental setup Figure 2. Engine brake thermal efficiency variation with engine torque for different fuels
21
Figure 3. Engine brake specific fuel consumption (bsfc) change with engine torque for different fuels. Figure 4. Engine air to fuel mass ratio variation with engine torque for different fuels. Figure 5. Variation of specific heat ratio of the cylinder contents during combustion with indicated power for different fuels. Figure 6. Maximum cylinder pressure change with net indicated power for different fuels. Figure 7. Rate of cylinder pressure change during engine low load operation Figure 8. Rate of cylinder pressure change during engine medium load operation Figure 9. Rate of cylinder pressure change during engine full load operation Figure 10. Net heat release rate during engine low load operation Figure 11. Net heat release rate during engine medium load operation Figure 12. Net heat release rate during engine full load operation Figure 13. Change of coefficient of variation with indicated power for different fuels.
22
Table(1)
Table 1. Properties of diesel vs. DEE Fuel property
Diesel
DEE
Density, kg/m3
849
713
Gross calorific value, MJ/kg
45.5
36.873
Auto-ignition temperature, oC [9] Cetane number [9]
210-350 150-160 50
125
Viscosity, mm /s [11]
3.25
0.23
Stoichiometric air to fuel ratio, kgair/kgfuel [11]
14.4
11.2
180-360
35
0
21.6
2
o
Boiling point, C [8] Oxygen content, % by mass [8]
Table(2)
Table 2. Engine specifications Item
value
No. of cylinders
1
Maximum power, kW
3.5 at 3600 rpm
Compression ratio
22
Bore, mm
69
Stroke, mm
62
Connecting rod length, mm 3
Engine capacity, cm
104 232
Table(3)
Table 3. Types of tested fuels Fuel
Acronym
Diesel
D100
95% diesel + 5% DEE ( %by volume)
D95
90% diesel+10% DEE (%by volume)
D90
85% diesel+ 15% DEE (%by volume)
D85
Table(4)
Table 4. Properties of tested fuels Fuel type
Gross calorific value, MJ/kg
Density @ 15 oC, kg/m3
Diesel (D100)
45.5
849
DEE
36.87
713
D95
45.13
842
D90
44.76
835
D85
44.39
828
Figure(1)
Figure(2)
35
30
Thermal efficiency, %
25
20
15
D100 D95 D90 D85
10
5 0
1
2
3
4
5 Engine torque, Nm
6
7
8
9
10
Figure(3)
1 D100 D95 D90 D85
0.9
0.8
bsfc, kg/kWh
0.7
0.6
0.5
0.4
0.3
0.2 0
1
2
3
4
5 Engine torque, Nm
6
7
8
9
10
Figure(4)
110 D100 D95 D90 D85
100
Air to fuel mass ratio
90
80
70
60
50
40
30 0
1
2
3
4
5 Engine torque, Nm
6
7
8
9
10
Figure(5)
1.34 D100 D90 D85
1.33
Sepcific heat ratio
1.32
1.31
1.3
1.29
1.28
1.27
1.26
0.8
1
1.2
1.4 Indicated power, kW
1.6
1.8
2
2.2
Figure(6)
82
D100 D90 D85
Maximum cylinder pressure, bar
80
78
76
74
72
70
68
0.8
1
1.2
1.4 Indicated power, kW
1.6
1.8
2
2.2
Figure(7)
8 D100 D90 D85 6
dp/dth,bar/deg
4
2
0
−2
−4 340
350
360
370 Crankangle,deg
380
390
400
Figure(8)
6 D100 D90 D85
5
4
dp/dth,bar/deg
3
2
1
0
−1
−2
−3 340
350
360
370
380 Crankangle,deg
390
400
410
Figure(9)
6 D100 D90 D85
5
4
dp/dth,bar/deg
3
2
1
0
−1
−2
−3 340
350
360
370
380 Crankangle,deg
390
400
410
420
Figure(10)
25 D100 D90 D85 20
dq/dth,J/deg
15
10
5
0
−5 340
350
360
370 Crankangle,deg
380
390
400
Figure(11)
25 D100 D90 D85 20
dq/dth,J/deg
15
10
5
0
−5 340
350
360
370
380 Crankangle,deg
390
400
410
Figure(12)
20 D100 D90 D85 15
dq/dth,J/deg
10
5
0
−5 340
350
360
370
380 Crankangle,deg
390
400
410
420
Figure(13)
10 D100 D90 D85
9
8
COV, %
7
6
5
4
3
2
1
0.8
1
1.2
1.4 Indicated power, kW
1.6
1.8
2
2.2
Highlights
DEE was tested as a fuel additive in different proportions in a diesel engine Using the DEE as a fuel additive improved engine performance significantly Engine thermal efficiency increased and engine fuel consumption decreased The maximum heat release rate and cylinder pressure generally increased Both the combustion duration and engine stability slightly reduced
23