1. Introduction
Nowadays, the study and development of alternative fuels are gaining more and more interest as one of the possible solutions for reducing greenhouse gas and pollutant emissions [
1,
2]. Different gaseous fuels, such as natural gas, shale gas, and hydrogen, are currently being investigated in order to avoid the formation of particulate matter [
3,
4,
5]. In the truck sector, diesel fuel can be substituted with natural gas due to the different advantages offered by this fuel; the high
ratio leads to a reduction in CO
emissions with respect to other fuels, and the opportunity to use lean mixtures reduces the nitrogen oxide emissions and improves the thermal efficiency. Furthermore, the conversion costs associated with the modification of existing diesel-fueled engines into spark-assisted natural-gas-fueled engines are relatively low. Through the use of a simplified after-treatment system, this approach is widely used in light- and heavy-duty engines to meet emission standards [
6,
7]. On the other hand, natural gas presents an important drawback related to its low power density. To this end, different solutions have been investigated over the years, including direct injection systems, lean combustion coupled with an advanced ignition system, and dual-fuel operation [
8,
9,
10,
11,
12,
13].
Within this context, the study and understanding of all processes that take place inside a combustion chamber become fundamental for achieving high efficiencies and low pollutant emissions. In light-duty engines, the thermodynamic conditions encountered in the combustion chamber are quite different from those found in passenger car engines; higher pressures and temperatures are expected due to the use of turbocharging and high compression ratios [
14]. Moreover, since most natural gas engines are derived from compression-ignition ones, the intake system is designed to generate a swirling motion inside the combustion chamber, even if tumbling is more desirable for turbulence generation during the compression stroke. For this reason, different piston bowl geometries [
15,
16,
17] and new combustion chamber layouts [
8,
18] are currently being examined.
To support engine design, CFD has become a powerful and efficient tool for studying the combustion process, from ignition to the burn-out phase. Over the years, different combustion models have been proposed, and the main challenges were expressed according to the correct description of the ignition stage, the laminar-to-turbulent flame transition, and suitable prediction of the flame wrinkling in different regimes to correctly model fully turbulent combustion. The main parameters that can be used to identify these flame conditions are the laminar flame speed
, the turbulence intensity
, the turbulence integral length
, and the flame thickness
[
19].
The most commonly used combustion models found in the literature are the coherent flamelet model (CFM) and the
G-equation. The former is based on transport equations for
c, the combustion progress variable, and
, the flame surface density [
20,
21], whereas the latter solves a transport equation for the non-reacting scalar
G, thus bypassing the need for a source-term closure and the problems associated with counter-gradient diffusion [
19,
22,
23,
24]. The use of these models for the simulation of spark-ignition engines has been successfully validated over the years [
23,
25,
26,
27,
28] by using simplified or more detailed ignition models [
29,
30] and both RANS and LES models for turbulence [
25,
31]. However, depending on their applications, different versions of the CFM have been proposed in the literature. When using this approach, the reaction rate tends to have an “Eddy-Breakup”-like expression under equilibrium conditions [
32]; thus, it fails to reproduce the flame behavior near walls in complex flows and large geometries [
33]. Furthermore, the need to tune at least two model constants without a general agreement on the formulations to use for flame surface density production and destruction terms [
34] strongly affects the predictive capability of this approach. Regarding the
G-equation model, its implementation in CFD code is not straightforward due to the different definitions of the
G field, and appropriate constraints should be used to guarantee geometrical consistency and avoid numerical instabilities [
35].
The main scope of this work is the development and validation of a multi-dimensional model to predict combustion in spark-ignition natural gas engines. The one-equation flame area model (FAM) proposed by Weller was used for this purpose due to its ability to describe the main combustion phases with a limited number of tuning constants. This model solves the transport equation of the combustion regress variable
b, while an algebraic expression of the flame wrinkling factor
is used for the reaction rate. The main advantages with respect to the CFM and the G-equation model are the choice of a regress variable that ensures numerical stability and the possibility of taking suitable expressions for
from the literature [
19,
36]. Ignition is described through a deposition model, whereas the laminar-to-turbulent transition is considered by means of a semi-empirical model derived from the work of Herweg and Maly [
37]. Simulations were carried out by using the Lib-ICE code developed by the ICE Group of Politecnico di Milano. Lib-ICE is based on the OpenFOAM software, and it was successfully applied by the authors for the simulation of IC engines in previous works [
8,
38,
39,
40,
41,
42].
The proposed approach was validated by using experimental data from a modern light-duty SI natural gas engine. First, three different operating conditions were simulated at various engine speeds and loads with the original piston bowl shape. The computed results were compared with experimental data on the in-cylinder pressure, heat release rate, and pollutant emissions ( and ). Then, the model was applied to study two other piston bowl geometries that were manufactured and tested to try to improve the combustion efficiency. Experimental validation was carried out by comparing the computed and experimental data on the in-cylinder pressure, heat release rate, and pollutant emissions. The simulations allowed us to understand the effects of piston bowl geometry on the combustion process.
2. Numerical Models
In this section, the main aspects related to the computational models used are presented. First, an overview of the mesh management for the gas exchange and the combustion phases is given. Then, the combustion model developed in this work on the basis of the one-equation Weller approach is described. In particular, all sub-models used to represent the flame from ignition to the turbulent combustion stage, together with the strategies adopted to predict the laminar flame speed and composition of the burnt mixture, are reported in detail.
The governing equations were solved by using a RANS approach in order to speed up the computational time, and the model was used for turbulence with the standard coefficients suggested in the literature.
During the simulations, two different approaches were used to manage the mesh during the gas exchange and combustion process:
Exhaust and intake phases: Multiple deforming grids were used to accommodate the motion of the piston and valves [
40];
Compression and combustion phases: Dynamic mesh layering was employed. Layers of cells were added or removed above the piston surface during its motion [
43,
44].
At the end of gas exchange simulations, the flow field at the IVC was mapped onto the grid used for the combustion simulation.
2.1. Flame Area Model
The proposed combustion model was founded on the one-equation flame area model described in Weller’s work [
33]. Based on the laminar flamelet assumption, it describes a flame’s development through the regress variable
b and the flame wrinkling factor
. The former represents the unburned gas fraction, so it is equal to 1 in the fresh mixture and 0 when the charge is completely burned. The regress variable’s transport equation is, therefore, described by:
where
and
are the density of the mixture and unburned mixture, respectively,
is the turbulent viscosity, and
is the unstrained laminar flame speed. The contribution of the ignition process is represented by
, while the first term on the RHS is the reaction rate due to turbulent flame propagation. The term
is the flame wrinkling factor, and it is defined as the ratio between the turbulent and the unstrained laminar flame speed
. The relation between the flame wrinkling factor and flame surface density
is described by:
One of the advantages of using Weller’s formulation is that Equation (
1) can be solved fully implicitly by exploiting differential operator properties. This ensures numerical stability and independence from the time step, both of which are very important when real geometries with complex grids are used.
2.2. Ignition Model
An initial distribution of the regress variable is required to start the flame propagation process. This task is carried out with a simplified deposition model, such as the one used in [
45], whereas more complex approaches validated by other authors [
28,
46] will be integrated into future work. Starting from a user-defined initial flame kernel diameter
and time interval
, an ignition source term is imposed in the transport equation in the cells for which the distance from the spark plug is less than
:
where
is a user-defined coefficient,
is the unburned gas density, and
is the ignition duration specified by the user.
2.3. Turbulent Combustion Model
Once the regress variable is initialized by the ignition model and the flame propagation process is started, a suitable expression for the flame wrinkling factor
that allows the flame front to develop from its initial laminar state to a fully turbulent flame is required, as shown in the first RHS term of Equation (
1). In this work,
is modeled according to authors’ previous work [
41].
The expression for the flame wrinkling factor under the equilibrium condition
is taken from the Peters formulation [
19]:
where
is the flame thickness, and it is defined as:
Here,
is the heat conductivity,
is the heat capacity, and
is the density, all of which are evaluated in the unburned gas. The values of the constants
,
, and
are reported in
Table 1.
2.4. Laminar Flame Speed Correlation
Laminar flame speeds are read from a lookup table in which the
values are stored as a function of the unburned temperatures, pressures, equivalence ratios, and EGR. This table is generated by a one-dimensional laminar flame speed solver operating under constant-pressure conditions. The fuel was assumed to be pure methane and the GRI mechanism was used in the calculation [
47].
2.5. Species Calculation
The chemical composition in each cell is calculated while knowing the mass fraction of the chemical species in the burned
and unburned
states and the regress variable
b:
Burned gas chemical composition
is computed in this work with two different methodologies. The first one is based on tabulated kinetics, and the approach was described in detail in [
48]. The chemical compositions and reaction rates are stored in a lookup table that is generated through constant-pressure homogeneous reactor calculations at different values of pressure, equivalence ratios, and unburned gas temperatures. To access the lookup table, the unburned gas enthalpy
transport equation is solved, which provides the temperature of the fresh charge
. Differently, the burned gas enthalpy
is computed from
, the mean cell value
h, and the regress variable:
Accordingly, the burned gas temperature is computed from and the composition .
The second approach is based on chemical equilibrium according to [
49]. The burned gas chemical composition
is, therefore, calculated at each time step with an iteration scheme, and the chemical system is composed of four elements (C, H, O, N) and ten reacting species (H
O, CO
, CO, O
, H
, N
, H, O, OH, NO).
3. Experimental Setup
The experimental setup was composed of a Schenck Dynas bench with a Horiba STARS automation system, as shown in
Figure 1. The natural gas fuel was composed, on average, of
methane on a molar basis, with small variations among the different CNG bottles. Then, the remaining molar shares consisted of ethane (
), higher hydrocarbons and hydrogen (
), and inert gases (
), such as nitrogen and carbon dioxide. The fuel mass flow rate was measured by using a Rheonik RHM 015 GNT Coriolis-type flow meter with an associated uncertainty of
.
All tests were performed by using conditioned air at 24
C and
relative humidity. The air mass flow rate was measured by using a hot-film anemometer (ABB Sensyflow FMT700-P) with an associated uncertainty of
.
Figure 2 presents the scheme of the test bench and its major measurement devices.
The experimental data were measured by using two different systems:
Crank-angle-based data were collected by using a Kistler Kibox triggered by a Kistler 2614A1 crank-angle sensor. Moreover, cylinder 1 was equipped with a piezoresistive Kistler 4007B sensor in the intake manifold and a water-cooled piezoresistive Kistler 4049B10 sensor in the exhaust manifold. The Kistler Kibox was used to control the spark timing and the knock index derived from the cylinder pressure.
Time-based information, such as pressure, temperature, mass flows, feedback signals from actuators, and data from the exhaust gas analyzer, were recorded by the test bench automation system. An anti-aliasing filter was used on each analog channel, and a frequency of 1 Hz was used to record 60 s of measuring time.
Figure 3 shows the scheme of all of the pressure, temperature, and lambda sensors mounted on the engine.
4. Simulated Conditions
To validate the proposed methodology for the simulation of the combustion process in an SI engine fueled with natural gas, full-cycle simulations of a four-cylinder turbocharged engine—whose main geometric data are reported in
Table 2—were performed. In particular, three different piston bowl geometries were studied on Empa test benches; the original piston was compared with two other geometries that, from preliminary considerations, should have given an increase in turbulence inside the combustion chamber and, thus, improved the combustion process. A star shaped piston was selected according to the work of Wohlgemuth et al. [
17], where a similar piston bowl geometry was investigated with the aim of improving the turbulence inside the combustion chamber at TDC. In particular, a star-like shape was chosen to disrupt the swirl charge motion due to the inner edges and to transform it into turbulence. The number of leaves was chosen as a compromise between a rugged geometry and higher heat losses. According to Wohlgemuth’s work, the modified piston bowl geometry was able to increase the turbulent kinetic energy at TDC by up to
. The tower piston was inspired by the activity of Heuser et al. [
16], where a geometry characterized by squish areas with cutout sections was considered, with the aim of converting the swirl motion into turbulent kinetic energy inside the combustion chamber. The results presented by Heuser showed that the mean turbulent kinetic energy was not highly affected by the proposed piston bowl geometry, whereas a zone of higher turbulence was found in the area underneath the combustion chamber’s ceiling, which was particularly relevant for early flame development.
Figure 4 illustrates the details of the three different piston bowl geometries.
Three different operating conditions were investigated, as reported in
Table 3: one low-load point running at 1000 rpm, one medium-load point running at 2200 rpm, and one full-load point running at 3500 rpm. These operating conditions were chosen in order to validate the proposed methodology under considerably different conditions without changing the model constants.
First, simulations of the gas exchange process were performed—starting from EVO and going until IVC—to study the effect of the piston bowl shape on the turbulence distribution during the intake process. The average in-cylinder cell size was 1.5 mm, and, during the gas exchange process, the total number of computational cells ranged from 1.5 to 1.7 million, depending on the piston bowl geometry considered (
Figure 5). The number of meshes required for one simulation varied from 80 to 90 depending on the piston bowl geometry and the operating conditions considered. Simulations were carried out on a 24-core machine (Intel(R) Xeon(R) CPU E5-2690 v4 processor with 2.60 GHz), and the CPU time varied from a minimum of 18 hours to a maximum of 2 days, depending on the operating conditions and on the considered mesh.
Figure 6 shows the computational grid used for the combustion simulations, whereas
Table 4 reports the main mesh information. Because the natural gas injection was not taken into consideration during the gas exchange phase, a uniform distribution of natural gas was imposed in the combustion mesh under stoichiometric conditions.
6. Conclusions and Future Development
The work presented in this paper was focused on the development and application of a combustion model for simulating spark-ignition engines operating with natural gas. The one-equation flame area model proposed by Weller was chosen due to its implicit formulation that ensures numerical stability and due to the limited number of tuning constants required.
The proposed approach was validated against a set of experimental data of a light-duty SI natural gas engine running at different engine speeds and loads. First, the model assessment was performed by using the original piston, and the results in terms of pressure, heat released, and pollutant emissions could be considered rather satisfactory for all of the tested conditions, even though some improvements are required, especially for high-load conditions. Then, two other piston bowl geometries were simulated under the same operating conditions to investigate the main differences in terms of turbulence generation and combustion evolution. The results were in rather good agreement with the experimental data, even though some discrepancies were visible, especially for low-load conditions. A more detailed description of the boundary conditions is probably required to improve the results. However, the proposed methodology was able to be successfully used for a preliminary investigation of the main advantages and disadvantages of changing the piston bowl geometry in a reasonable amount of time thanks to the use of the RANS approach. Concerning the effects of the new piston bowl geometries, no important improvements in terms of combustion efficiency were found, which was probably due to the very low () or null increase in turbulent kinetic energy during the intake and compression stroke and due to the greater heat losses () associated with the increased surface area of the new piston bowls.
Future works will focus their attention on alternative approaches to predicting the flame wrinkle factor and introducing crevice geometry in order to correctly predict HC emissions. Moreover, the proposed approach will also be used for the simulation of lean mixtures with more complex combustion systems, such as active prechambers and different low-carbon fuels, such as hydrogen.