2.3.3. Rejected Heat

The thermodynamic model used Equation (16) to predict the heat transfer rate from the combustion gases to the combustion chamber walls.

$$\frac{dQ\_r}{d\theta} = \frac{h\_{wall} \bullet A\_{wall} \bullet (T - T\_{wall})}{2 \bullet \pi \bullet N} \tag{16}$$

where *Twall* and *Awall* are the temperature and the wall surface area. Besides, *hwall* is the heat transfer coefficient, calculated via the correlation proposed by Woschni [36] as given by Equation (17):

$$h\_{wall} = 3.26 \bullet (w \bullet P)^{0.8} \bullet T^{-0.55} \bullet b^{-0.2} \tag{17}$$

where *b* accounts for the internal diameter of the combustion chamber and *w* represents the average velocity of the combustion gases. The latter is calculated via Equation (18).

$$w = k\_1 \bullet (2 \bullet N \bullet S\_t) + k\_2 \bullet \frac{T\_o \bullet V\_d \bullet (P - P\_{\text{ff}})}{P\_o \bullet V\_o} \tag{18}$$

Subscript "*o*" relates the initial state for volume (*Vo*), pressure (*Po*) and temperature (*To*). *P* and *Pm* are the actual and mean pressure in the combustion chamber. *k*1 and *k*2 are model constants, whose values are defined as 2.30 and 3.25 × <sup>10</sup>−3, respectively. Finally, *St* represents the engine stroke.

#### 2.3.4. Combustion Chamber Volume

Here we define the combustion chamber volumetric characteristics, which involve different volume contributions that incorporate operational and geometrical patterns to define the instantaneous volumetric displacement within each cycle [37]. Equation (19) gives the instantaneous volume of the combustion chamber.

$$V = V\_{mv} + V\_{disp} + \bullet V\_{dp} + \bullet V\_{inf} + \bullet V\_{c} \tag{19}$$

The term *Vmv*, calculated via Equation (20), represents the free space encountered once the piston reaches the top-dead center (TDC).

$$V\_{mv} = \frac{\pi \bullet D^2}{4} \bullet \left[\frac{2 \bullet L\_{cr}}{r\_c - 1}\right] \tag{20}$$

where *rc* and *D* are the compression ratio and the internal diameter of the piston, respectively. *Lcr* represents crankshaft length and *Vdisp* the volume displaced by the connecting rod-crank mechanism. The latter is predicted using Equation (21).

$$V\_{disp} = \frac{\pi \bullet D^2}{4} \bullet \left[L\_{cr} + L\_{rod} - R\_{\bar{y}} \bullet \theta\right] \tag{21}$$

*Lrod* is the longitude of the crankshaft and *Ry* represents the vertical position of the piston. Consequently, the volumes denoted as Δ*Vdp* and Δ*Vinf* are calculated using Equations (22) and (23), respectively. The former represents the variation of the instantaneous volume induced by pressure-deformation effects imposed by the combustion gases. The latter accounts for the volume related to the inertial forces of the connecting rod-crank shaft mechanism.

$$
\Delta V\_{dp} = \frac{\pi \bullet D^2 \bullet L\_{rod}}{4 \bullet A\_c} \bullet \left(\frac{k\_{def}}{E\_s}\right) \bullet \left(P \bullet A\_p\right) \tag{22}
$$

$$
\Delta V\_{inf} = \frac{\pi \bullet D^2 \bullet L\_{rod}}{4 \bullet A\_c} \bullet \left(\frac{k\_{d\varepsilon f}}{E\_s}\right) \bullet \left(m \bullet a\_p\right) \tag{23}
$$

where *kdef* , *Es*, *Ac*, *Ap* and *ap* are the deformation constant, the elastic modulus of steel, connecting rod critical area, the piston cross-sectional area, and piston acceleration, respectively.

Finally, Δ*Vc* is calculated according to Equation (24) and represents the variation of the instantaneous volume produced by the clearances in the combustion chamber [37].

$$
\Delta V\_{\mathfrak{c}} = -\frac{\pi \bullet D^2}{4} \bullet \sum\_{i=1}^{2} (e\_i \bullet \sin \varphi\_i \bullet \cos \alpha\_i) \tag{24}
$$

where e represents the eccentricity between journal and bearing, measured along their centerline. Similarly, *ϕ* relates to the angle of rotation, and *α* represents the angle between the connecting rod and the piston.

#### 2.3.5. Energy Distribution and Emissions Processing

.

The engine inlet heat energy ( *Qint*) is defined based on the fuel mode, which can be entirely liquid using pure gasoline or oxygenated fuel blends with ABE as indicated in Equation (25), or alternately liquid-gaseous fuel, which constitutes the dual-fuel operation mode given by Equation (26).

.

.

$$
\dot{Q}\_{int} = \dot{m}\_{fuel} \bullet LHV\_f \tag{25}
$$

$$
\dot{Q}\_{\rm int} = \dot{m}\_{fuel} \bullet LHV\_{fuel} + \dot{m}\_{HHO} \bullet LHV\_{HHO} \tag{26}
$$

where the subscripts fuel refers to the fuel mixture and HHO represents the hydroxy gas. Similarly, the model defines the exhaust gas energy as follows:

.

$$Q\_{\rm cxh} = \dot{m}\_{\rm cxh} \bullet C\_{p,\rm cxh} \bullet T\_{\rm cxh} \tag{27}$$

where *Cp*,*exh* represents the specific heat capacity of the combustion gases at constant pressure. On the other hand, the mechanical efficiency of the engine according to the fuel mode is defined as follows.

Liquid fuel standalone operation:

$$
\eta\_{m\text{cch}} = \frac{PW}{\dot{m}\_{fuel} \bullet LHV\_{fuel}} \bullet 100\tag{28}
$$

Dual-fuel mode:

$$\eta\_{mech} = \frac{PW}{\dot{m}\_{fuel} \bullet LHV\_{fuel} + \dot{m}\_{HHO} \bullet LHV\_{HHO}} \bullet 100\tag{29}$$

where *PW* is the power output of the engine calculated according to Equation (30).

$$PW = \frac{2 \bullet \pi \bullet \omega \bullet T\_r}{60 \bullet 1000} \tag{30}$$

where *Tr* is the torque condition, and *ω* refers to the engine angular speed. Lastly, this section concludes with the calculation of the unit conversion of the overall emissions.

On the other hand, the measuring instrumentation for emissions levels is commonly reported in ppm and %vol, which are the default characteristics of gas analyzers. However, it is necessary to apply further processing to display these results according to the international standards (i.g. European legislation) that describe pollutants in terms of g • km−<sup>1</sup> for light-dutty and passenger vehicles and g • kWh−<sup>1</sup> for heavy-duty vehicles [38]. Therefore, the study converts the output data from the gas analyzers into g • kWh−<sup>1</sup> according to the empirical correlations proposed by Heseding and Daskalopoulos [39]. The relevance of the emissions above processing is that they assist in relating emissions and fuel metrics, facilitating comparison. The correlation for emissions processing is based on the following formulation:

$$EP\_{\bar{i}} = EV\_{\bar{i},dry} \bullet \left(\frac{M\_{\bar{i}}}{M\_{cxh,dry}} \bullet k\_d\right) = EV\_{\bar{i},wet} \bullet \left(\frac{M\_{\bar{i}}}{M\_{cxh,wet}} \bullet k\_w\right) \tag{31}$$

where *EPi*, *EVi*,*dry* and *EVi*,*wet* represents the pollutant mass in the power unit (g • kWh−1), exhaust emissions on a dry basis and wet basis, respectively. The term *Mi* accounts for the molecular mass, while *Mexh*,*<sup>d</sup>* and *Mexh*,*<sup>w</sup>* relates to the molecular mass of exhaust emissions on a dry and wet basis, respectively. Finally, the terms *kdry* and *kwet* are empirical constants with a value of 3.873 g • kWh−<sup>1</sup> and 4.160 g • kWh−1, respectively, and relate the power unit and the exhaust emissions on a dry basis and wet basis, respectively. The conversion of the three primary pollutants treated in experimental evaluation is described in Equation (32) to (34) [39].

$$\text{CO} \left[ \frac{\text{g}}{kWh} \right] = 3.591 \bullet 10^{-3} \bullet \text{CO} (\% \, vol) \tag{32}$$

$$\text{KOx}\left[\frac{\mathcal{g}}{kWh}\right] = 6.636 \bullet 10^{-3} \bullet \text{NOx}(ppm) \tag{33}$$

$$\left[HC\left[\frac{\mathcal{S}}{kWh}\right] = 2.002 \bullet 10^{-3} \bullet HC(ppm) \tag{34}$$

#### **3. Results and Discussion**

## *3.1. Cylinder Pressure*

The experimental evaluation of the dual-fuel operation in the SI engine begins with the characterization of the pressure gradients developed within the combustion chamber. This parameter provides a clear perspective of the fuel mode performance while relating the appropriate mixing interaction between the base fuel (gasoline/ABE), hydroxy gas (gaseous fuel), and air. Figure 3 shows the overall behavior of the in-cylinder pressure at different load conditions. Notice that the standalone gasoline operation has been set as the baseline fuel for comparison purposes.

**Figure 3.** *Cont.*

**Figure 3.** Influence of fuel on cylinder pressure for an engine load of (**a**) 50%, (**b**) 75%, and (**c**) 100%.

According to the results, the maximum in-cylinder pressure within the engine operating range is achieved by the ABE 5 + HHO with values between 60–135 bar. This result is consistent with the magnification of the laminar flame speed of this blend. Results in Figure 3 assist in the evaluation of the combustion phasing from the different fuel blends. Based on Figure 3a, all the blends present a retarded phasing between 0.2◦–0.6◦ compared to the baseline fuel, which implies that a low engine load inhibits the fast combustion for the alternative fuels. The overall pattern encountered in the phasing is consistent with related investigations of ABE at different blend ratios [12]. Next, in Figure 3b, all the tested fuels feature a relatively similar phasing condition. However, at high engine loads (Figure 3c), the enhancement on the laminar flame speed from the oxygenated fuels becomes evident as the combustion phasing is advanced, ranging from 0.8◦ to 1.6◦ compared to the baseline, which facilitates the completion of combustion and faster peak pressures achievement even before reaching 360◦. The latter is a direct indication of a significant improvement in combustion efficiency and thermal performance.

The pressure curve features an increasing trend as the engine load increases since more fuel is mixed to meet the power demand. Notably, increasing the fuel substitution with ABE (ABE 10) limited the pressure produced during combustion between 30–70 bar, which can be associated with the lower calorific value of this blend. This result agrees with similar investigations related to diesel engine combustion phenomena while operating with biodiesel blends [40]. Another contributor to the reduced pressure of ABE-based blends could be associated with a higher-octane rating that extends the initiation delay and reduces the laminar flame speed. In contrast, hydroxy doping enables a significant improvement in the combustion pressure in both blends (ABE 5/10) up to 75%. This pattern can be explained considering that incorporating hydroxy in the engine facilitates a homogeneous air-fuel mixture. The positive effects on the gasoline octane rating by incorporating HHO in the intake air system can be mentioned as another contributor to the enhanced combustion performance since the compression ratio is maximized. It should be noted that the overall trend of the pressure curves after combustion features a sharp decrement which guarantees that the knocking condition is not reached. On average, the combustion pressure developed in the standalone gasoline operation is higher than that of ABE 5, ABE 10, and ABE 10 + HHO by 12%, 18%, and 24%, respectively. Contrarily, the ABE 5 + HHO shows an enhanced pressure range compared to commercial gasoline between 10–15% for the engine load ranges analyzed.

#### *3.2. Heat Release Rate (HRR)*

The *HRR* relates to the fuel conversion efficiency since it shows how much chemical energy is transformed into thermal energy. Figure 4 depicts the *HRR* at different engine load conditions as a function of the crankshaft angle.

**Figure 4.** *Cont.*

**Figure 4.** Influence of fuel on heat release rate for (**a**) 50%, (**b**) 75%, (**c**) 100% of the engine load.

According to the results, the general pattern in all the load conditions states that ABE 5 + HHO features the highest heat release from the fuel blends, followed by gasoline, ABE 5, ABE 10 + HHO, and ABE 10. It can be observed that a higher engine load promotes the increase of heat release in the combustion chamber, which implies extreme conditions and maximum chemical energy conversion. The average reduction of the *HRR* for the ABE 5 and ABE 10 was 3.5% and 6.78%, respectively, compared to gasoline as the baseline fuel. The latter can be explained considering the higher viscosity of these blends that promote a slower combustion process, thus limiting heat release.

HHO enrichment increases heat release in all the blends, which implies integrating gaseous fuel mitigates the cooling effect associated with utilizing oxygenates (ABE). The higher calorific value and subsequent enhancement in the laminar flame speed resulting from HHO doping can be mentioned as contributors to the enhanced behavior in the heat release that offset the decrement experienced by ABE blends when compared to pure gasoline. It is worth mentioning that the fuel chemical structure supports the enhanced behavior from HHO doping since both hydrogen and oxygen coexist in the air/fuel mixture, whereas gasoline consists of hydrocarbon molecules [41]. Therefore, the gaseous fuel incorporation promotes an improved combustion performance due to the direct interaction of the diatomic molecules that suppress the ignition delay. In this sense, HHO doping also fosters the massive bond-breaking trend of the gasoline molecules, thus facilitating the heat release rate, the laminar flame speed, and subsequently improving combustion efficiency.

#### *3.3. Combustion Chamber Temperature*

Figure 5 shows the average temperature of the combustion chamber for the tested fuels. This parameter indicates the ability of a blend for combustion phasing. Notice that the study only presents the temperature distribution for a full load rate, representing the critical condition from the analyzed cases since it holds the highest heat release rate and maximum pressure peaks.

**Figure 5.** In-cylinder temperature at a full load rate.

According to the results, the maximum temperature is achieved by ABE 5 + HHO, followed by the gasoline, ABE 5, ABE 10 + HHO, and ABE 10. Interestingly, the temperature after combustion presents a reverse trend concerning the blends, which can be associated with the higher heat release rate during combustion, minimizing the temperature in this stage. The maximum temperature reaches a value of 2229 ◦C for ABE 5 + HHO, supporting the energetic contribution derived from the dual-fuel operation. Moreover, as the ABE content escalates, the combustion temperature drops between 20–45%. This pattern can be explained considering the lower heating value and intensified latent heat of the ABE blends. Notably, increasing the latent heat promotes a temperature drop in the intake stage, resulting in a lower temperature at the end of the compression stage. Moreover, incorporating hydroxy gas in the blends intensifies the temperature due to the higher hydrogen and oxygen content that stimulates the chemical energy conversion of the air/fuel mixture [42]. On average, the combustion chamber temperature of ABE 5 and ABE 10 decreased by 6.21% and 12.23% compared to the gasoline fuel. In contrast, HHO enrichment increases the temperature of ABE 5 and ABE 10 by up to 394 ◦C and 341 ◦C, respectively.
