**2. Research methodology**

The aim of the present chapter is to develop the one-dimensional model of four-stroke port fuel injection (PFI) gasoline engine for predicting the effect of methanol-gasoline (M0–M50) and ethanol-gasoline (E0–E50) addition to gasoline on the exhaust emissions and performance of gasoline engine. For this, simulation of gasoline SI engine (calibrated) as the basic operating condition and the laminar burning velocity correlations of methanol-gasoline and ethanol-gasoline blends for calculating the changed combustion duration was used. The engine power, specific fuel consumption, and exhaust emissions were compared and discussed [22, 23].

#### **2.1. Simulation setup**

The one-dimensional SI engine model is created by using the AVL BOOST software and has been employed to examine the performance and emissions working on gasoline, ethanolgasoline, and methanol-gasoline blends.

In **Figure 1**, PFIE symbolizes the engine, while C1–C4 is the number of cylinders of the SI engine. The cylinders are the main element in this model, because they have many very important parameters to settle: the internal geometry, bore, stroke, connecting rod, length and compression ratio, as well as the piston pin offset and the mean crankcase pressure. The measuring points are marked with MP1–MP18. PL1–PL4 symbolizes the plenum. System boundary stands for SB1 and SB2. CL1 represents the cleaner. R1–R10 stands for flow restrictions. CAT1 symbolizes catalyst and fuel injectors—I1–I4. The flow pipes are numbered 1–34.

The calibrated gasoline engine model was described by Iliev [23], and its layout is shown in **Figure 1** with engine specification shown in **Table 1**.

**Table 2** presents a comparison between the properties of gasoline, ethanol, and methanol. As shown in **Table 2**, compared with gasoline and ethanol, methanol has a higher elemental oxygen content and a lower heating value, molecular weight, elemental carbon, hydrogen content, and stoichiometric air/fuel ratio (AFR).

#### **2.2. Combustion model description**

In this research, two-zone model of Vibe was chosen for the combustion simulation and analysis. The combustion chamber was divided into two regions: unburned gas region and burned gas regions [17]. For the burned charge and unburned charge, the first law of thermodynamics is applied:

$$\frac{dm\_b\mu\_b}{d\alpha} = -p\_\varepsilon \frac{dV\_b}{d\alpha} + \frac{dQ\_F}{d\alpha} - \sum \frac{dQ\_{\theta b}}{d\alpha} + h\_\nu \frac{dm\_b}{d\alpha} - h\_{\theta 0, b} \frac{dm\_{\theta b, b}}{d\alpha} \tag{1}$$

 *dmu <sup>u</sup>* \_\_\_\_\_\_*<sup>u</sup> <sup>d</sup>* <sup>=</sup> <sup>−</sup>*pc dV*\_\_\_\_*<sup>u</sup> <sup>d</sup>* <sup>−</sup> <sup>∑</sup>\_\_\_\_\_ *dQWu <sup>d</sup>* <sup>−</sup> *hu* \_\_\_\_ *dmB <sup>d</sup>* <sup>−</sup> *hBB*,*<sup>u</sup>* \_\_\_\_\_ *dmBB*,*<sup>u</sup> <sup>d</sup>* (2)

where *dmu*

work, *dQ*\_\_\_\_*<sup>F</sup>*

represents the change of the internal energy in the cylinder, *pc*

flow from the unburned to the burned zone due to the conversion of a fresh charge to combus-

Moreover, the sum of the zone volumes must be equal to the cylinder volume, and the sum of

*Vb* + *Vu* = *V* (4)

*<sup>d</sup>* <sup>+</sup> *dV*\_\_\_\_*<sup>u</sup> <sup>d</sup>* <sup>=</sup> \_\_\_ *dV*

*da* is the wall heat loses, *hu*

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143

*da* stands for the fuel heat input, *dQ*\_\_\_\_*<sup>W</sup>*

**Figure 1.** Schematic of the gasoline PFI engine model.

*dV*\_\_\_\_*<sup>b</sup>*

tion products. The heat flux between the two zones is neglected. *hBB*

the volume changes must be equal to the cylinder volume change:

due to blow-by, **u** and **b** in the subscript are unburned and burned gas.

\_*dV*\_\_

*da* represents the enthalpy

*da* is the enthalpy flow

*dm*\_\_\_\_*<sup>b</sup>*

*dm*\_\_\_\_\_*BB*

*<sup>d</sup>* (3)

*da* is the piston

Comparison of Ethanol and Methanol Blending with Gasoline Using Engine Simulation http://dx.doi.org/10.5772/intechopen.81776 143

**Figure 1.** Schematic of the gasoline PFI engine model.

**2. Research methodology**

142 Biofuels - Challenges and opportunities

discussed [22, 23].

**2.1. Simulation setup**

gasoline, and methanol-gasoline blends.

The aim of the present chapter is to develop the one-dimensional model of four-stroke port fuel injection (PFI) gasoline engine for predicting the effect of methanol-gasoline (M0–M50) and ethanol-gasoline (E0–E50) addition to gasoline on the exhaust emissions and performance of gasoline engine. For this, simulation of gasoline SI engine (calibrated) as the basic operating condition and the laminar burning velocity correlations of methanol-gasoline and ethanol-gasoline blends for calculating the changed combustion duration was used. The engine power, specific fuel consumption, and exhaust emissions were compared and

The one-dimensional SI engine model is created by using the AVL BOOST software and has been employed to examine the performance and emissions working on gasoline, ethanol-

In **Figure 1**, PFIE symbolizes the engine, while C1–C4 is the number of cylinders of the SI engine. The cylinders are the main element in this model, because they have many very important parameters to settle: the internal geometry, bore, stroke, connecting rod, length and compression ratio, as well as the piston pin offset and the mean crankcase pressure. The measuring points are marked with MP1–MP18. PL1–PL4 symbolizes the plenum. System boundary stands for SB1 and SB2. CL1 represents the cleaner. R1–R10 stands for flow restrictions. CAT1 symbol-

The calibrated gasoline engine model was described by Iliev [23], and its layout is shown in

**Table 2** presents a comparison between the properties of gasoline, ethanol, and methanol. As shown in **Table 2**, compared with gasoline and ethanol, methanol has a higher elemental oxygen content and a lower heating value, molecular weight, elemental carbon, hydrogen

In this research, two-zone model of Vibe was chosen for the combustion simulation and analysis. The combustion chamber was divided into two regions: unburned gas region and burned gas regions [17]. For the burned charge and unburned charge, the first law of thermodynamics

> \_\_\_\_ *dmBd* <sup>−</sup> *hBB*,*<sup>u</sup>*

\_\_\_\_\_ *dmBB*,*<sup>u</sup>*

*<sup>d</sup>* (2)

(1)

izes catalyst and fuel injectors—I1–I4. The flow pipes are numbered 1–34.

**Figure 1** with engine specification shown in **Table 1**.

content, and stoichiometric air/fuel ratio (AFR).

*<sup>d</sup>* <sup>=</sup> <sup>−</sup>*pc*

*dV*\_\_\_\_*<sup>u</sup> <sup>d</sup>* <sup>−</sup> <sup>∑</sup>\_\_\_\_\_ *dQWu <sup>d</sup>* <sup>−</sup> *hu*

**2.2. Combustion model description**

*dmu <sup>u</sup>* \_\_\_\_\_\_*<sup>u</sup>*

is applied:

where *dmu* represents the change of the internal energy in the cylinder, *pc* \_*dV*\_\_ *da* is the piston work, *dQ*\_\_\_\_*<sup>F</sup> da* stands for the fuel heat input, *dQ*\_\_\_\_*<sup>W</sup> da* is the wall heat loses, *hu dm*\_\_\_\_*<sup>b</sup> da* represents the enthalpy flow from the unburned to the burned zone due to the conversion of a fresh charge to combustion products. The heat flux between the two zones is neglected. *hBB dm*\_\_\_\_\_*BB da* is the enthalpy flow due to blow-by, **u** and **b** in the subscript are unburned and burned gas.

Moreover, the sum of the zone volumes must be equal to the cylinder volume, and the sum of the volume changes must be equal to the cylinder volume change:

$$\frac{dV\_b}{da} + \frac{dV\_u}{da} = \frac{dV}{da} \tag{3}$$

$$V\_b + V\_u = V\tag{4}$$


The amount of burned mixture at each time setup is obtained from the Vibe function. For all other terms, for instance, wall heat losses, etc., models similar to the single-zone models with

In AVL BOOST, the model of formation on NOx is based on AVL List Gmbh [24], which incorporates the Zeldovich mechanism [25]. The rate of NOx production was obtained using

In the above equation, *CPPM* represents post-processing multiplier, *CKM* denotes kinetic mul-

*rCO* = *CConst*(*r*<sup>1</sup> + *r*2).(1 − *α*) (6)

where *<sup>α</sup>* <sup>=</sup> *<sup>C</sup>*\_\_\_\_\_\_ *CO*.*act*

The unburned HC has different sources. A complete description of HC formation still cannot be given, and the achievement of a reliable model within a thermodynamic approach is definitely prevented by the fundamental assumptions and the requirement of reduced computational times. Still, a phenomenological model which accounts for the main formation mechanisms and is able to capture the HC trends as function of the engine operating parameter may be proposed.

**1.** During the intake and compression stroke, fuel vapor is absorbed into the oil layer and deposits on the cylinder walls. The following desorption occurs when the cylinder pressure decreases during the expansion stroke and complete combustion cannot take place

**2.** A fraction of the charge enters the crevice volumes and is not burned since the flame

**3.** Occasional complete misfire or partial burning takes place when combustion quality is

**4.** Quench layers on the combustion chamber wall which are left as the flame extinguishes

The following important sources of unburned HC can be identified in SI engines [21]:

*CCO*.*equ* .

(

*<sup>r</sup>* \_\_\_\_\_\_\_ <sup>1</sup> 1 + *αAK*<sup>2</sup>

Comparison of Ethanol and Methanol Blending with Gasoline Using Engine Simulation

<sup>+</sup> *<sup>r</sup>* \_\_\_\_\_\_ <sup>4</sup>

*i*

*i*

<sup>1</sup> <sup>+</sup> *AK*4). (5)

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145

represents reaction rates of

represents reaction rates

an appropriate distribution on the two zones are used [24].

**2.3. A description of exhaust emission model**

*rNO* = *CPPM CKM*(2, 0).(1 − *α*<sup>2</sup>).

, *AK*<sup>4</sup> <sup>=</sup> *<sup>r</sup>* \_\_\_\_4 *r* <sup>5</sup> + *r* 6 .

tiplier, *C* stands for molar concentration in equilibrium, and *r*

The NOx formation model in AVL Boost is based on Onorati et al. [26]:

In Eq. (6), *C* represents molar concentration in equilibrium and *r*

, *AK*<sup>2</sup> <sup>=</sup> *<sup>r</sup>* \_\_\_\_1 *r* <sup>2</sup> + *r* 3

Eq. (5):

where *<sup>α</sup>* <sup>=</sup> *<sup>C</sup>*

\_\_\_\_\_\_ *NO*.*act CNO*.*equ* . \_\_\_\_ 1 *CKM*

Zeldovich mechanism.

based on the model.

anymore.

poor.

quenches at the entrance.

prior to reaching the walls.

**Table 1.** Engine specification.


**Table 2.** Comparison of fuel properties.

The amount of burned mixture at each time setup is obtained from the Vibe function. For all other terms, for instance, wall heat losses, etc., models similar to the single-zone models with an appropriate distribution on the two zones are used [24].

#### **2.3. A description of exhaust emission model**

In AVL BOOST, the model of formation on NOx is based on AVL List Gmbh [24], which incorporates the Zeldovich mechanism [25]. The rate of NOx production was obtained using Eq. (5):

$$r\_{\rm MO} = C\_{\rm PPM} C\_{\rm KM} \langle 2,0 \rangle . (1 - \alpha^2) . \left( \frac{r\_1}{1 + \alpha A K\_2} + \frac{r\_4}{1 + A K\_4} \right). \tag{5}$$

where *<sup>α</sup>* <sup>=</sup> *<sup>C</sup>* \_\_\_\_\_\_ *NO*.*act CNO*.*equ* . \_\_\_\_ 1 *CKM* , *AK*<sup>2</sup> <sup>=</sup> *<sup>r</sup>* \_\_\_\_1 *r* <sup>2</sup> + *r* 3 , *AK*<sup>4</sup> <sup>=</sup> *<sup>r</sup>* \_\_\_\_4 *r* <sup>5</sup> + *r* 6 .

**Properties Gasoline Methanol Ethanol**

Molecular weight 111.21 32.04 46.07 Oxygen content (wt%) — 49.93 34.73 Carbon content (wt%) 86.3 37.5 52.2 Hydrogen content (wt%) 24.8 12.5 13.1 Stoichiometric AFR 14.5 6.43 8.94 Lower heating value (MJ/kg) 44.3 20 27 Heat of evaporation (kJ/kg) 305 1.178 840 Research octane number 96.5 112 111 Motor octane number 87.2 91 92 Vapor pressure (psi at 37.7 OC) 4.5 4.6 2

H15 CH3

)

)

) 0.737 0.792 0.785

Normal boiling point (OC) 38–204 64 78 Autoignition temperature (OC) 246–280 470 365

OH C2

H5 OH

Chemical formula C8

**Table 1.** Engine specification.

**Engine parameters Value** Bore 86 (mm) Stroke 86 (mm) Compression ratio 10.5

144 Biofuels - Challenges and opportunities

Connection rod length 143.5 (mm)

Intake valve open 20 BTDC (deg) Intake valve close 70 ABDC (deg) Exhaust valve open 50 BBDC (deg) Exhaust valve close 30 ATDC (deg) Piston surface area 5809 (mm<sup>2</sup>

Cylinder surface area 7550 (mm<sup>2</sup>

Number of stroke 4

Number of cylinder 4 Piston pin offset 0 (mm) Displacement 2000 (cc)

Destiny (g/cm<sup>3</sup>

**Table 2.** Comparison of fuel properties.

In the above equation, *CPPM* represents post-processing multiplier, *CKM* denotes kinetic multiplier, *C* stands for molar concentration in equilibrium, and *r i* represents reaction rates of Zeldovich mechanism.

The NOx formation model in AVL Boost is based on Onorati et al. [26]:

$$
\sigma\_{\rm CO} = \mathcal{C}\_{\rm Coat} (r\_1 + r\_2) \cdot (1 - \alpha) \tag{6}
$$

$$
\text{where } \alpha = \frac{\mathcal{C}\_{\rm COat}}{\mathcal{C}\_{\rm COat}} \text{ .}
$$

In Eq. (6), *C* represents molar concentration in equilibrium and *r i* represents reaction rates based on the model.

The unburned HC has different sources. A complete description of HC formation still cannot be given, and the achievement of a reliable model within a thermodynamic approach is definitely prevented by the fundamental assumptions and the requirement of reduced computational times. Still, a phenomenological model which accounts for the main formation mechanisms and is able to capture the HC trends as function of the engine operating parameter may be proposed. The following important sources of unburned HC can be identified in SI engines [21]:


**5.** The flow of fuel vapor into the exhaust system during valve overlap in gasoline engines.

The first two mechanisms and in particular the crevice formation are considered to be the most important and need to be accounted for in a thermodynamic model. Partial burn and quench layer effect cannot be physically described in a quasi-dimensional approach, but may be included by adopting tunable semiempirical correlations.

The formation of unburned HC in the crevices is described by assuming that the pressure in the cylinder and in the crevices is the same and that the temperature of the mass in the crevice volumes is equal to the piston temperature.

The mass in the crevices at any time is described by Eq. (7):

$$m\_{cenive} = \frac{pV\_{cenire}M}{RT\_{piston}}\tag{7}$$

**3. Result and discussion**

**3.1. Engine performance characteristics**

tion is obtained at E50 ethanol-gasoline blend.

**Figure 2.** Influence of ethanol-gasoline blended fuels on brake power.

output is obtained.

The present research focused on the performance and emission characteristics of the methanol and ethanol-gasoline blends. Various concentrations of the blends 0% methanol (ethanol) M0 (E0), 5% methanol (ethanol) M5 (E5), 10% methanol (ethanol) M10 (E10), 20% methanol (ethanol) M20 (E20), 30% methanol (ethanol) M30 (E30), 50% methanol (ethanol) M50 (E50),

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147

The results of the brake power and specific fuel consumption for ethanol-gasoline blended

The brake power is one of the important factors that determine the performance of an engine. The variation of brake power with speed was obtained at full load conditions for E5, E10, E20, E30, E50, and pure gasoline E0. The ethanol content in the blended fuel increased, and the brake power decreased for all engine speeds. The gasoline brake power was higher than E5–E50 for all engine speeds. The ethanol's heat of evaporation is higher in comparison to gasoline fuel, providing air-fuel charge cooling and increasing the density of the charge. The blended fuel causes the equivalence ratio of blend approaches to stoichiometric condition which can lead to a better combustion. However, the ethanol heating value is lower compared to gasoline, and it can neutralize the previous positive effects. Consequently, a lower power

**Figure 3** shows the changes of the BSFC for ethanol-gasoline blends under various engine speeds. The figure shows that the BSFC increased as the ethanol percentage increased. Heating value and stoichiometric air-fuel ratio are the smallest for these two fuels, which means that for specific air-fuel equivalence ratio, more fuel is needed. The highest specific fuel consump-

and 85% methanol (ethanol) M85 (E85) by volume were analyzed.

fuels at different engine speeds are shown on **Figures 2** and **3**.

In Eq. (7), *mcrevice* represents the mass of unburned charge in the crevice, *p* denotes cylinder pressure, *Vcrevice* stands for total crevice volume, *M* represents unburned molecular weight, *Tpiston* is the temperature of the piston, and *R* denotes gas constant.

The second important source of HC is the presence of lubricating oil in the fuel or on the walls of the combustion chamber. During the compression stroke, the fuel vapor pressure increases so, by Henry's law, absorption occurs even if the oil was saturated during the intake. During combustion the concentration of fuel vapor in the burned gases goes to zero so the absorbed fuel vapor will desorb from the liquid oil into the burned gases. Fuel solubility is a positive function of the molecular weight, so the oil layer contributed to HC emissions depending on the different solubility of individual hydrocarbons in the lubricating oil.

The assumptions made in the development of the HC absorption/desorption are the following:


The radial distribution of the fuel mass fraction in the oil film can be determined by solving the diffusion Eq. (8):

$$\frac{\partial w\_f}{\partial t} - D \frac{\partial^2 w\_f}{\partial r^2} = 0 \tag{8}$$

In Eq. (8), *wF* represents fuel's mass fraction in the oil film, *t* is the time, *r* stands for radial position in the oil film (distance from the wall), and *D* is relative (fuel-oil) diffusion coefficient.
