ORIGINAL ARTICLE

Thermodynamic Evaluation of Using Ethanol-Gasoline Blends in Spark Ignition Engine

Ing. Yarian Reyes Suarez^{1  *} 

Dr.C. Yanoy Morejon Mesa¹ 

Est. Abel Hernández Herranz¹ 

¹Universidad Agraria de La Habana, Facultad de Ciencias Técnicas, Centro de Mecanización Agropecuaria (CEMA), San José de las Lajas, Mayabeque, Cuba.

ABSTRACT

The present investigation was developed in the Engines Laboratory that belongs to the Faculty of Technical Sciences, in the Agrarian University of the Havana (UNAH). For its development a spark ignition engine, Jacto model, was used. The objective of this investigation was to evaluate thermodynamically the use of blends ethanol-gasoline. To achieve the outlined objective, different percentages of ethanol were added to the gasoline (E-10%; E-15%; E-20%; E-25%), to carry out an incomplete combustion with coefficient of excess of air $\propto$ =0,85 and an complete combustion with coefficient of excess of air $\propto$ =1,5, for the different percentages of ethanol. Among the main results obtained it was evidenced that when using pure gasoline, the necessary air-fuel rate to achieve the combustion, is bigger than when blends are used with ethanol, because ethanol contains oxygen, and that impoverishes the blend and guarantees a better quality in the combustion process. On the other hand, it was verified that for E-10 and E-15 blends, a better environmental and energy behaviour was obtained due to the increment of the octane rating or antiexplosive capacity of the fuel blends used which improves the quality of the combustion, although decreases the energy power during the explosion (detonation).

Keywords: Combustion; internal energy; fuel blends; air-fuel rate

INTRODUCTION

In order to produce optimum combustion in an internal combustion engine, fuel vaporization is necessary (^{Kozlov et al., 2019}). The vaporization temperatures of pure ethanol are very high, due to its low vapor pressure and its high latent vaporization heat, so it is necessary to implement some kind of ignition aid system. The simplest way is to mix it with gasoline in different percentages according to the needs (^{Mantilla et al., 2016}).

Biofuels respond to government policies on energy security, which are intended to replace partially the limited fossil fuels to reduce the threat to the environment by escaping emissions and global warming. The use of alcohols, which are considered important forms of biofuels, produced from biomass (for example, alcohol, biofuel, bio-kerosene, H₂, etc.); blended with gasoline as motor fuel, has been a subject of scientific research since 1980 (^{Gravalos et al., 2011}; ^{Guarieiro and Guarieiro, 2013}).

Among the different alcohols, ethanol and methanol are recognized as the most suitable renewable biofuels for spark ignition internal combustion engines (^{Kamboj and Karimi, 2014}).

The use of ethanol-gasoline blends with a low percentage of ethanol, such as E-5 (5% ethanol, 95% gasoline), can be used in any type of gasoline vehicle, without any mechanical modification (^{Kheiralla et al., 2017}).

According to ^{Kozak (2019)}, ethyl alcohol has been used for many years as an additive to regular gasoline, so that in Poland and most other European countries, the market for vehicles fuelled with pure ethanol or blends with a high ethanol content is not developed, having a significant market share.

Technologies for the production of ethanol from non-food feedstocks, mainly from lignocellulosic complexes, have not yet had a significant impact on this fuel production market. Although ethanol continues being a very important component for spark ignition engines.

The ethanol has a great impact due to its physical and chemical properties, since it contains oxygen, which propitiates obvious effects in the oxidation of the particles. In the case of the pressure in an internal combustion engine with ethanol mixed with gasoline, it is 36% higher than that reached with pure gasoline. The combustion process is shortened with the increase of ethanol and makes it approaching to an ideal combustion to constant volume. The increase of ethanol proportion can suppress the formation, accumulation and emission of particles. With 100% of ethanol, the increase of the engine regime can reduce the accumulation of particles up to 72%, without affections with the increase of the speed, whereas with the increase of the gasoline in the blend the accumulation of particles increases abruptly.

On the other hand, the advance of direct injection synchronizing can effectively reduce the total number of particles and the effects of early direct injection on the decrease of particles are more obvious at high temperatures in engine speed. At low speed, the delay in ignition time, causes the particles to decrease (^{Sun et al., 2019}).

The combined injection engine, fully qualified for ethanol and gasoline injection, can also control the ratio of the two fuels in real time and obtain good engine performance to meet energy saving and emission reduction requirements. For the production of such an engine, a set of fuel injection devices is added (^{Gutiérrez, 2013}).

METHODS

Theoretical Foundations for the Combustion Process Analysis

The theoretical foundations for the analysis of the combustion process considered were those stated by ^{Anojin and Sajarov (1970)}, ^{Gurevych and Sorokin (1978)}, ^{Pancratov (1979)}, ^{Kralob and Antonov (1980)}, ^{Vzorov et al. (1981)} and ^{Jovarj (1982)}. They established the following:

The theoretical amount of air required to oxidize a fuel is determined by the following expression:

${\mathrm{L}}_{\mathrm{o}}=\frac{1}{\mathrm{0,21}}\left(\frac{\mathrm{C}}{12}+\frac{\mathrm{H}}{4}-\frac{{\mathrm{O}}_{\mathrm{o}}}{32}\right)$ (1)

The average fuel elemental components can be taken: for gasolines: C=0,85; H= 0,15; O_o =0 and for diesel fuel: C=0,86; H=0,13; O_o =0,01.

The quantity of fresh load moles for gasoline engines:

$M_1=\mathrm{\alpha }\mathrm{*}{L}_o+\frac{1}{{\mu }_c}$ (2)

where: μc: molecular mass of the fuel, for gasolines μc=110…120 kg/kmol.

For diesel, the amount of fresh charge is calculated by the expression:

$\mathrm{M}1=\mathrm{\alpha }\mathrm{*}\mathrm{L}\mathrm{o}$ (3)

Quantity of combustion products:

If $\alpha <1$

${\mathrm{M}}_2=\frac{\mathrm{C}}{12}+\frac{\mathrm{H}}{2}+\mathrm{0,79}\mathrm{*}\mathrm{\alpha }\mathrm{*}{\mathrm{L}}_{\mathrm{o}}$ (4)

If $\alpha \ge 1$ then

${\mathrm{M}}_2={\left({\mathrm{M}}_2\right)}_{\mathrm{\alpha }=1}=1+\mathrm{J}$ (5)

where:

${\left({\mathrm{M}}_2\right)}_{\mathrm{\alpha }=1}=\frac{\mathrm{C}}{12}+\frac{\mathrm{H}}{2}+\mathrm{0,79}\mathrm{*}{\mathrm{L}}_{\mathrm{o}}$ (6)

In addition, there is an amount of air that is in excess, which is equal to:

$\left(\mathrm{\alpha }-1\right){\mathrm{L}}_0=\mathrm{J}$ (7)

The theoretical coefficient of molecular variation:

${\mathrm{\mu }}_{\mathrm{o}}=\frac{{\mathrm{M}}_2}{{\mathrm{M}}_1}$ (8)

The real coefficient of molecular variation:

$\beta =\frac{M_2+M_r}{M_1+M_r}=\frac{M_2+M_1*{\sigma }_r}{M_1+(1+{\sigma }_r)}=\frac{{\mathrm{\mu }}_{\mathrm{o}}+{\mathrm{\sigma }}_{\mathrm{r}}}{1+{\sigma }_r}$ (9)

The magnitude of $\beta$ oscillates for gasoline engines $\beta =\mathrm{1,02}\dots \mathrm{1,12}$ , diesel engines $\beta =\mathrm{1,01}\dots \mathrm{1,06}$ .

The lower combustion heat of the fuel is formed: for gasoline engines ${ H}_u=44 000 kJ/k_g$ , for diesel $H_u=42 500 kJ/k_g$ , and for gasohol (Flex fuel) ${ H}_u=26\mathrm{ }279 kJ/k_g$ .

For gasoline engines working with rich blends $\alpha >1$ , the loss of heat as a result of incomplete combustion is determined.

$∆{\mathrm{H}}_{\mathrm{u}}=119\mathrm{ }600\left(1-\mathrm{\alpha }\right)*L_o; kJ/k_g$ (10)

If $\alpha \ge 1$ , the previous equation does not apply. $$

The internal energy of combustion products is determined by the expression:

$\frac{1}{\beta }\left[\xi \frac{\left(H_u-∆H_u\right)}{M_1\left(1+{\sigma }_r\right)}+\frac{\left(U_c-{\sigma }_r{*U_c}^{\text{'}\text{'}}\right)}{1+{\sigma }_r}\right]=U_z$ (11)

where: ${\sigma }_r$ : Coefficient of waste gases: for gasoline engines ${\sigma }_r$ =0,06…0,12 ; for diesel engines ${\sigma }_r$ =0,02…0,06; $\xi$ : Heat utilisation coefficient for different engines at nominal speed: for gasoline engines $\xi =\mathrm{0,85}\dots \mathrm{0,95}$ ; for diesel engines $\xi =\mathrm{0,65}\dots \mathrm{0,85}$ . $U_c$ : Internal gas energy (the one from the air can be taken) for the temperature at the end of compression in degrees Celsius ( $℃$ ); $U_c{}^{\text{'}\text{'}}$ : Internal energy of combustion products at the critical temperature of the products de $t_c$ .

Exhaust Process

The parameters of the exhaust process $P_r$ y $T_r$ , are taken at the beginning of the calculation of the admission process. This process consists of filling the cylinder with fresh load. The pressure of the environment is considered $P_o=0,1MPa$ , while the temperature of the environment is determined at the place where the experiment takes place.

The residual gas pressure, depending on the engine type, is calculated using the following equation:

${\mathrm{P}}_{\mathrm{r}}=\left(\mathrm{1,05}\dots \mathrm{1,25}\right)\mathrm{*}{\mathrm{P}}_{\mathrm{o}}$ (12)

The temperature of the waste gases is taken depending on the type of engine. Considering that in gasoline engines its value oscillates between $T_r=900\dots 1 100°K$ , while for diesel engines its value lies between the ranges $T_r=700\dots 900°K$ .

Depending on the type of engine, the heating temperature of the load would be between $∆T=10\dots 30°K$ .

To check the correct selection of the value of $P_r$ y $T_r$ , the following equation is used:

${\mathrm{T}}_{\mathrm{r}}=\frac{{\mathrm{T}}_{\mathrm{b}}}{\sqrt[3]{\frac{{\mathrm{P}}_{\mathrm{b}}}{{\mathrm{P}}_{\mathrm{r}}}}}$ (13)

The selected value of $T_r$ and the one calculated by the Equation 13, must not be greater than 10%, otherwise the thermal calculation must be corrected.

The approximate values of $T_b$ y $P_b$ are listed in Table 1.

TABLE 1 Values of temperatures and final expansion pressures 

Engine type	$T_b,°K$	$P_b;MPa(kgf/{cm}^2)$
Gasoline	2 200…1 700	0,34…0,50 (3,4…5,0)
Diesel	1 800…1 200	0,2…0,4 (2…4)

RESULTS AND DISCUSSION

Characterization of Experimental Conditions

The experiments took place in the Laboratory of Motors of the Faculty of Technical Sciences, at the Agrarian University of Havana (UNAH), using a single-cylinder engine model JACTO.

The analysis of the combustion process in terms of ethanol-gasoline and pure gasoline blends was carried out in April 2019. Before the combustion, the percentage of ethanol to be added was defined (10%; 15%; 20% y 25%), and the type of blend (rich mixture and lean mixture). Hydrated ethanol with 95% purity and regular gasoline category B-85 were used as fuel. The engine used has an effective power of 1,2 kW; a rotation frequency of 580 rev/min and a cylinder capacity of 34 cm³

Thermodynamic Analysis of the Combustion Process

Considering what was proposed in the materials and methods, referring to the influence of the air composition in the chemical equations to be obtained, when fuel blends were used, the following analyses were carried out. Air-fuel and fuel-air ratio, internal energy, mass and molar fraction of the combustion products, the quantity of gases emitted to the environment and the economic analysis.

In the particular case of the air-fuel and fuel-air ratio, it was determined for excess air coefficients of 0.85 and 1.5, for each of the ethanol-gasoline blends and pure gasoline supplied to the engine.

In the case of the rich blend, i.e. with an excess air coefficient equivalent to 0.85, the values of the air-fuel and fuel-air ratios are shown in the Table 2.

TABLE 2 Analysis of r_a/c y r_c/a for the combustion process for rich blend ( $\propto =1,5$ ) 

Blends	ra/c, kg(air)/kg(fuel)	rc/a, kg(fuel)/kg(air)
E-0	16,049	0,062
E-10	12,599	0,079
E-15	13,048	0,077
E-20	12,592	0,079
E-25	12,587	0,079

As it can be seen in the table above, for the blend with E-0, the air-fuel ratio required to achieve combustion reaches a value of 16.049 kg (air)/kg (fuel). This value is higher than the values obtained in the rest of the blends, reaching a maximum difference of 3.462 kg (air)/kg (fuel) with respect to the blend with E-25 and a minimum difference of 3.001 kg (air)/kg (fuel) with respect to the blend with E-15.

When analysing the fuel-air ratio represented in Table 2, it can be seen that, for blend with E-10, E-20 and E-25, equal values are reached, these being equivalent to 0.079 kg (fuel)/kg (air). It is observed that this value is higher than that in the blend with E-0 by 0.017 kg (fuel)/kg (air), however, when a blend with E-15 is used, a difference of 0.015 kg (fuel)/kg (air) with respect to that of E-0 is evident.

In turn, the internal energy of the combustion products for each of the blend analyzed was determined by means of Expression 11. It is evident that the maximum value of internal energy in the combustion products is reached in the blend with E-0, which is given by the caloric power that is given off when pure gasoline only is used (Table 3).

TABLE 3 Internal energy of combustion products for rich blend ( $\propto =1,5$ ) 

Blend /Thermodynamic parameters	M1, kg/kmol	M2, kg/kmol	M0	β	Uz, kJ/kmol
E-0	36,970	28,943	0,783	0,806	22 516,084
E-10	37,723	30,718	0,814	0,834	16 499,621
E-15	37,661	31,074	0,825	0,844	17 346,729
E-20	37,737	30,732	0,814	0,834	16 393,280
E-25	37,737	30,497	0,808	0,829	17 667,364

As it can be seen, in the table above, when using the E-0 blend, a higher internal energy is obtained than when the remaining mixtures were used. A value of 22 516,084 kJ/kmol is reached and a maximum difference of 6 122,804 kJ/kmol is obtained with respect to the mixture with E-25, and a minimum difference of 4 848,72 kJ/kmol is obtained when using a mixture with E-20.

The decrease in the internal energy of the combustion products is mainly due to the increase in octane rating, i.e. the anti-explosive capacity of the fuel or fuel blend used, an aspect that improves the quality of combustion, although energy power is reduced during the explosion (detonation).

When using a light mixture with an excess air coefficient equivalent to 1.5, it is observed that the values of the air-fuel and fuel-air ratio differ from those obtained when using a rich mixture, which is given by the excess air supplied to the combustion chamber. These values are shown in Table 4.

TABLE 4 Analysis of r_a/c y r_c/a for the combustion process for lean blend ( $\propto =1,5$ ) 

Blend	ra/c, kg(air)/kg(fuel)	rc/a, kg(fuel)/kg(air)
E-0	28,329	0,035
E-10	23,171	0,043
E-15	22,225	0,045
E-20	22,212	0,045
E-25	22,205	0,045

According to Table 4, the values of blends E-15, E-20 and E-25 have similar values, with the E-25 mixture reaching the minimum value with 22.205 kg(air)/kg(fuel), showing a maximum difference with respect to the E-0 mixture of 6.124 kg(air)/kg(fuel). While for the E-10 blend, a minimum difference of 5,158 kg(air)/kg(fuel) is observed with respect to the E-10 blend and a maximum difference of 6,124 kg(air)/kg(fuel) is observed.

When analysing the fuel-air ratio, it is evident that when using blends E-15, E-20 and E-25, values equal to 0.045 kg(fuel)/kg(air) are obtained. That evidences a maximum difference with respect to the E-0 blend of 0.010 kg(fuel)/kg(air), while the E-10 blend reaches a value of 0.043 kg(fuel)/kg(air), evidencing a minimum difference with respect to the E-0 blend of 0.008 kg(fuel)/kg(air).

Using Expression 11, the internal energy of the combustion products was determined for each of the blends analyzed, evidencing that the maximum value of internal energy in the combustion products is reached in blends with E-0 (Table 5).

TABLE 5 Internal energy of combustion products for lean blend ( $\propto =1,5$ ) 

Blend /Thermodynamic parameters	M1, kg/kmol	M2, kg/kmol	M0	β	Uz, kJ/kmol
E-0	36,371	33,370	0,908	0,913	21 928,776
E-10	36,372	33,070	0,909	0,914	16 315,053
E-15	36,802	33,318	0,905	0,910	16 305,665
E-20	36,802	33,307	0,905	0,910	16 305,665
E-25	36,800	32,740	0,890	0,896	16 557,925

In Table 5, it is observed that when using E-0 blends, the internal energy is 21 928.776 kJ/kmol, reaching a value higher than the other blends studied, showing a minimum difference with respect to the E-25 blend of 5 370.851 kJ/kmol, while the maximum difference is 5 623.111 kJ/kmol, with respect to the E-15 and E-20 blends.

To summarise, in the combustion analysed, the highest internal energy value of the combustion products is achieved when the E-0 blend is used, independently of the excess air coefficient. However, it is valid to point out that in the majority of experiments it is evident that the best values with respect to the energy behaviour of the products obtained are obtained when blends E-10 and E-15 are used. This behaviour is due to the increase in octane rating, that is, to the anti-knock capacity of the fuel or fuel blend used, an aspect that improves the quality of combustion, although energy power is reduced during the explosion (detonation).

In the research developed by ^{Pikūnas et al. (2003)}, it is reflected that when ethanol is added, the heat released by the ethanol-gasoline blend decreases, however, it is observed that the octane rating increases, aspects that coincide with the results obtained in this research.

CONCLUSIONS

• The theoretical foundations were validated by evaluating the thermodynamic behaviour of the use of ethanol-gasoline blends in two-stroke spark ignition engines.

• When using pure gasoline, the air-fuel ratio needed to achieve combustion is higher than when using ethanol blends. This is due to the fact that ethanol contains oxygen, an aspect that impoverishes the blend and guarantees a better quality in the combustion process.

• The E-10 and E-15 blends experienced the best environmental and energy performance, which is due to the increase in octane rating, that is, the anti-knock capacity of the fuel blend used, an aspect that improves the quality of combustion, although energy power is reduced during the explosion (detonation).

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⁵The mention of trademarks of specific equipment, instruments or materials is for identification purposes, there being no promotional commitment in relation to them, neither by the authors nor by the publisher.

Received: September 10, 2019; Accepted: March 13, 2020

^*Author for correspondence: Yarian Reyes Suarez, e-mail: [e-mail]

Yarian Reyes Suarez, Especialista, Universidad Agraria de La Habana, Facultad de Ciencias Técnicas, Centro de Mecanización Agropecuaria (CEMA) San José de las Lajas, Mayabeque, Cuba, e-mail: [e-mail]

Yanoy Morejon Mesa, Profesor Titular, Universidad Agraria de La Habana, Facultad de Ciencias Técnicas, Centro de Mecanización Agropecuaria (CEMA) San José de las Lajas, Mayabeque, Cuba, e-mail: [e-mail]

Abel Hernández Herranz, Estudiante de Ingeniería Agrícola, Universidad Agraria de La Habana, Facultad de Ciencias Técnicas, San José de las Lajas, Mayabeque, Cuba, e-mail: [e-mail]

The authors of this work declare no conflict of interests.