Effect of a Taper Intake Port on the Combustion Characteristics of a Small-Scale Rotary Engine

Abstract

Taper intake ports are effective in improving the charging efficiency of small-scale rotary engines (REs), but it is unclear how their structural parameters affect the in-cylinder flow field and combustion characteristics. For this reason, the effects of the diameter-length ratio (D/L) of an intake port on the in-cylinder flow field and combustion characteristics of a small-scale RE were numerically investigated by utilizing a three-dimensional computational fluid dynamics (CFD) model. The results showed that the in-cylinder pressure of the RE did not follow a simple single-directional trend with the D/L of the intake port, but it was divided into three levels, where the peak in-cylinder pressure was at its maximum at the D/L of 0.6 and at its minimum at the D/L of 0.8. The gas flows in the intake port with different values of the D/L were all unidirectional, and they made a difference in the vortexes formed on the leading side of the combustion chamber of the RE, which was the main factor affecting the in-cylinder combustion performance. The vortexes formed on the leading side of the combustion chamber with D/L = 0.6 were maintained for a long period of time, thus promoting the propagation of flame and advancing the center of gravity of combustion. So, the heat release rate and combustion efficiency in the cylinder were increased at the price of a larger increment in nitrogen oxide formation.

1. Introduction

In comparison with reciprocating piston engines, rotary engines (REs) possess many advantages, such as a simple structure, low vibration, low noise, high speed, high power-to-weight ratio, small frontal area, etc. [1,2]. These advantages make REs promising for applications in small unmanned aerial vehicles, electric vehicle range extenders, armored military vehicles, small ships, etc. [3,4]. However, REs also possess some drawbacks, such as poor sealing and high fuel consumption [5,6]. In particular, the elongated rotor chamber and the unidirectional flow field in the cylinder make it difficult for the flame to propagate to the trailing side of the combustion chamber, resulting in severely inadequate combustion [7]. To improve the combustion performance of REs, researchers have carried out some work from the perspectives of structural optimization, ignition scheme improvement, multi-fuel mixing, intake process optimization, etc. [8,9,10]. Unlike a traditional piston engine, an RE does not have intake valves and exhaust valves. Instead, the mixture enters the combustion chamber directly from the intake port and collides with the rotor recess wall. This collision affects the distribution of the in-cylinder flow field and the intensity of the turbulent kinetic energy (TKE), which, in turn, affects the efficiency of in-cylinder combustion. Therefore, the intake process of the RE is more significantly affected by the intake port structure.

The intake process of an RE is affected by a variety of factors of the intake port structure, mainly including the intake port’s cross-sectional shape, its location, its structural size parameters (diameter, length, etc.), etc. Fan et al. [11] numerically investigated the effects of intake shape, including a round intake, rectangular intake, regular trapezoidal intake, and inverted trapezoidal intake, on the flow field in the cylinder of an RE, and they found that the inverted trapezoidal intake was the most conducive to improving the volume coefficient and holding a swirl structure. Ji et al. [12] studied the influences of a peripheral-ported intake, side-ported intake, and compound intake on the in-cylinder flow field, the flame propagation, and the emission formation of a gasoline RE. The results showed that the peripheral-ported intake had the fastest flame propagation and the highest combustion rate. At the moment of opening of the exhaust valve, the NOx emission was the lowest for the side-ported intake, and the CO emission was the lowest for the compound intake. Taskiran et al. [13] analyzed the effects of multiple side-ported intakes and exhaust ports on the flow field and combustion process in the combustion chamber of an RE and showed that the two side-ported intakes had the best charge coefficients and the fullest combustion of the mixtures. Jeng et al. [14] studied the influence of the length and diameter of an intake pipe and exhaust pipe with an equal cross-sectional area on the performance of an RE and concluded that the shorter the intake pipe and the longer the exhaust pipe, the higher the output power. These studies were carried out to assess the influence of intakes with an equal cross-section on the flow field and combustion performance, and this contributed to the improvement of the intake process of REs, but it did not involve studies of variable-section intake ports, such as a taper intake port. A taper intake port improved intake efficiency by increasing the differential pressure [15,16]. In general, a small-scale RE uses a taper intake port in order to achieve a high intake efficiency. The length and diameter of a taper intake port directly affect the intake resistance and intake pressure difference, which affects the mass flow rate and velocity of air entering the cylinder, and then influences the flow field distribution and combustion characteristics in the cylinder. However, due to the lack of studies on taper intake ports, it is unclear how the changes in intake charge and intake velocity caused by changes in the diameter and length of a taper intake port synthetically affect the in-cylinder flow field and combustion characteristics of a small-scaled RE.

In this paper, a three-dimensional computational fluid dynamics (CFD) model of a peripheral-intake RE was established in commercial software based on a reduced chemical reaction mechanism and the applicable turbulence model, and its accuracy was verified through a comparison with experimental results. Then, the influences of the diameter–length ratio of the taper intake port on the flow field, TKE, flame propagation, and combustion process inside the combustion chamber of the smal-scale RE were analyzed, which, in turn, provided theoretical guidelines for the design of the intake port of a small-scale RE.

2. Numerical Procedure

2.1. Engine Geometry and Computational Domain

The object of this study is a peripheral-ported-intake single-rotor engine with a dual spark plug ignition mode, whose model schematic and model parameters are shown in Figure 1 and

Table 1. Specifications of the small-scale RE.

Specifications	Value
Generating radius	52.5 mm
Eccentricity	7.5 mm
Width of rotor	33.75 mm
Compression ratio	10.5
Displacement	70 cc
Ignition source	two spark plugs
Ignition timing	30°EA BTDC
Intake timing	80°EA ATDC, 40°EA ABDC
Exhaust timing	60°EA BBDC, 70°EA ATDC

Note: EA is the eccentric angle, BTDC is before top dead center, BBDC is before bottom dead center, ATDC is after top dead center, and ABDC is after bottom dead center.

, respectively.

Chamber III of the tested RE was selected as the calculation object in this study due to the identical working process of the three combustion chambers, which corresponded to a single rotor of the RE, and due to the consideration of the symmetrical similarity of the construction and working principle of the RE. The starting position of the calculation was 720°EA BTDC. In addition, to eliminate the influence of backflow on the exhaust flow, the exhaust port was extended in the calculation (as shown in the right diagram of Figure 1).

2.2. Mathematical Models and Boundary Conditions

The flow field in the cylinder of an RE includes cyclonic flow and unidirectional flow, making it a complex turbulent flow. Among several commonly used turbulence models, some researchers have concluded that the RNG k-$\epsilon$ model can more accurately characterize the flow field inside the cylinder of an RE by comparing simulation results with experimental results [13,17]. Therefore, the RNG k-$\epsilon$ model was also chosen in this study.

The fuel used in this RE is RON92 gasoline, which includes 92% (v/v) iso-octane and 8% (v/v) n-heptane. The reduced mechanism of 47 species and 142 reactions developed by Ra et al. [18] was chosen in order to achieve combustion of the gasoline. This mechanism can be used to accurately calculate the combustion reaction of this gasoline, and it has been verified in various basic combustion and internal combustion engine combustion with an accuracy of the calculation results that satisfies the requirements [19,20].

The SAGE combustion model is able to calculate the detailed chemical reaction mechanism in turbulent flow [21]. In this study, the fuel was injected into the intake port and mixed with air; then, it entered the combustion chamber of the RE for compression and combustion. The gasoline combustion mechanism used in this study is a multi-step reaction. Consequently, the SAGE combustion model is well suited for the current simulation’s calculations [22]. In order to assess the emissions of the small-scale RE, the reactant generation models of both “Thermal NO” and “Prompt NO” were included in the simulation’s calculations [4]. In addition, the ignition model used in this study was based on the application of a virtual spherical flame kernel with a diameter of 1 mm and an ignition energy of 50 mJ to replace the actual spark plug in triggering the combustion reaction in the combustion chamber.

Pressure boundary conditions were used for the boundaries of the inlet and the outlet. Since the intake method of the small-scale RE in this study was naturally aspirated, the pressures of the inlet and the outlet were set to the ambient pressure of 0.101 MPa, and the temperatures were set to the ambient temperature of 300 K. All walls were set as no-slip boundaries, and the temperatures of the cylinder walls and rotor walls were set to 400 K. The temperatures of the intake and exhaust port walls were set to 300 and 550 K, respectively.

In addition, the numerical simulation implemented in this study made the following assumptions: On the one hand, the flow field of the combustion chamber was assumed to be a transient and compressible flow; on the other hand, the inlet flow was assumed to be a uniform flow.

2.3. Grid Generation

In order to obtain a mesh that was adapted to the changes in the combustion chamber structure, a homemade file describing the rotor motion was introduced into the CONVERGE software to automatically update the combustion chamber mesh with the rotor motion. The element type of a hexahedral mesh was selected in this study. In addition, to achieve better accuracy in the calculation results, the rotor walls, the local region of the injection beam, and the local regions near the spark plug position were treated with a fixed refined mesh. Adaptive mesh refinement (AMR) with temperature and velocity was set in the combustion domain, and the relationship between the grid size and the refinement level was as follows:

$$dx=\frac{d{x}_{base}}{2^n}$$

(1)

where $dx$ is the grid size, $d{x}_{base}$ represents the basic grid size, and n denotes the refinement level.

To eliminate the influence of the grid size on the computational results, in our previous research, we conducted an irrelevance analysis of several grid size schemes, and we found that 2.5 mm + AMR was the best grid scheme for balancing the computational efficiency and accuracy [23]. Consequently, the grid size scheme of 2.5 mm + AMR was chosen for this study. The total number of cells in the grid was around 5000~35,000 per chamber.

2.4. Research Approach

A taper intake port was used to replace the equal-section intake port of the original engine (as shown in the right diagram of Figure 1). A schematic diagram of the taper intake port studied in this paper is shown in Figure 2, where the section of $L_0$ is the injector installation position, and the parameters of this section were not changed in the designed test. The final calculation schemes that were developed are shown in

Table 2. Design of the diameter–length ratios of different intake ports.

Case	D/L	Case	D/L
1	0.2	5	0.6
2	0.3	6	0.7
3	0.4	7	0.8
4	0.5	8	0.9

, where D/L = 0.2 indicates that the tapering inlet diameter (D) has the same value as that of the outlet diameter (d).

2.5. Model Validation

To ensure the simulation accuracy of the 3-D CFD simulation model of the RE established in this study, the parameter settings of the model need to be verified. Under operating conditions with an equivalent ratio of 1.25, a speed of 4500 rpm, and a wide-open throttle, a comparison was made between the simulation results and experimental results for the in-cylinder pressure, as shown in Figure 3. As can be seen from the figure, the in-cylinder pressure curve of the numerical simulation was slightly higher than the curve of the experimental results at the early stage of combustion (near the TDC). The main reasons for this were: (1) There were only two fuel components used in the simulation, which did not fully replace the multiple components of actual gasoline. (2) Leakage through the rotor tip was neglected in the simulation model. (3) The simplified chemical reaction mechanism of the alternative fuel that was adopted did not sufficiently consider the quenching phenomenon. Fortunately, the comparison showed a good agreement between the experiments and simulation with a deviation of less than 5%. The deviation values were within acceptable limits and on par with similar reference simulations from around the globe. Therefore, the 3-D CFD model established in this study can accurately predict the combustion process of a small-scale RE.

3. Results and Discussion

3.1. Effect of D/L on In-Cylinder Pressure

Figure 4 shows the variation in the in-cylinder pressure according to the eccentric angle with different D/L parameters. It can be seen that when D/L was 0.6, the peak in-cylinder pressure was at its maximum, and when D/L was 0.8, the peak in-cylinder pressure was at its minimum. In addition, the in-cylinder pressure was divided into three levels in all calculated tests. The maximum in-cylinder pressure was found for D/L values of 0.2 and 0.6, the second highest was found for D/L values of 0.3, 0.5, 0.7, and 0.9, and the lowest was found for D/L values of 0.4 and 0.8. As illustrated in Figure 5, the peak in-cylinder pressure first decreased and then increased with the D/L parameter in a cyclic pattern, which was not a simple linear trend in a single direction. The eccentric angle corresponding to the peak pressure varied in proportion to the peak pressure. The above phenomenon was mainly caused by the influence of the change in the size of the taper intake port on the intake efficiency and the in-cylinder flow field of the small-scale RE.

Figure 6 shows the results for the volumetric coefficient with different D/L parameters. From the figure, it can be seen that the volumetric coefficient showed a slightly increasing trend with the increase in D/L, except for the sudden increment in the volumetric coefficient for the intake port structure scheme with D/L = 0.4. This was mainly because as the D/L increased, the pressure difference between the constriction and expansion sections of the intake port increased, thus boosting the flow velocity of the mixtures in the constriction section and increasing the amount of air intake. However, the previous results showed that the peak in-cylinder pressure first decreased and then increases with the D/L parameter, rather than a simple change trend in a single direction, which indicated that the variation in the volumetric coefficient caused by the change in the D/L was not the main reason for the effect on the performance of the RE.

To find out how the parameters of the taper intake port affected the engine performance, representative D/L parameters of 0.3, 0.6, and 0.8 at three pressure levels were selected for a subsequent analysis of the in-cylinder flow field and combustion characteristics of the small-scale RE.

3.2. Effect of D/L on the In-Cylinder Flow Field Characteristics

The change in the parameters of the taper intake port inevitably caused variations in the intake charge and intake flow velocity, which, in turn, affected the evolution of the in-cylinder flow field. Figure 7 displays the distribution of the velocity field and streamlines in the central plane of the combustion chamber at different moments for different values of D/L. It can be clearly seen that the flow field in the intake port was unidirectional at all times, without backflows or vortexes, which was the same as the results that were obtained by other researchers [16]. At the early stage of the intake stroke (500°EA BTDC), the mixtures from the intake port impinged on the rotor wall and moved in the direction of rotation, and there was no vortex phenomenon in the cylinder and a lower flow velocity in the intake port. This was because at the initial stage of the intake stroke, the smaller volume of the combustion chamber limited the development of vortexes, and the pressure difference between the combustion chamber and the intake port was minor, resulting in a low flow velocity in the intake port. At the median stage of the intake stroke (410°EA BTDC), one vortex was formed on the leading side and one on the trailing side of the combustion chamber. However, the vortex on the leading side of the combustion chamber for D/L = 0.8 was in the initial formation stage, and its intensity was lower than that of the vortex for D/L values of 0.3 and 0.6. In addition, as the D/L increased from 0.3 to 0.8, the velocity at the entrance of the combustion chamber was gradually enhanced. This was due to an increment in the contraction ratio at the throat of the taper intake port as the D/L increased. At the late stage of the intake stroke (360°EA BTDC), the vortex on the leading side of the combustion chamber with a D/L of 0.3 started to disappear, while the vortex on the leading side of the combustion chamber with D/L values of 0.6 and 0.8 showed an increasing trend in comparison with the median stage of the intake stroke. This was because the increase in the flow velocity at the combustion chamber’s entry caused by the rise of the D/L boosted the effect of the in-cylinder mixtures on the rotor walls. This increased the intensity of the mixture flow after impacting the rotor walls and enhanced the intensity and duration of the vortex on the leading side of the combustion chamber. From the results given above, it can be seen that the flow velocity in the intake port with a D/L of 0.6 enabled the vortex on the leading side of the combustion chamber to form early and disappear late, indicating that its vortex duration was longer, which was conducive to the formation of homogeneous mixtures in the combustion chamber. In fact, similar results were obtained by other researchers for the above rules of vortex evolution in the intake stroke [11,13].

Figure 8 depicts the variation curves of the TKE according to the eccentric angle with D/L values of 0.3, 0.6, and 0.8. It can be concluded from Figure 8 that at the intake stroke (600°EA~250°EA BTDC), the TKE in the cylinder for the D/L values of 0.6 and 0.8 was basically the same, while the TKE for D/L = 0.3 was higher than that of both of the other D/L values. This was because the smaller flow velocity in the intake port with D/L = 0.3 enabled vortexes on the leading side of the combustion chamber to form early, resulting in higher TKE in the cylinder at the early stage. However, during the whole compression stroke (250°EA BTDC~0°EA), the TKE for D/L = 0.8 was lower than that of the other two intake ports. This was because the vortexes on the leading side of the combustion chamber for D/L = 0.8 formed late and were maintained for a short time, resulting in a lower TKE in the cylinder at the later stage. In addition, the TKE in the cylinder during the compression stroke was essentially the same for the D/L values of 0.3 and 0.6, but after the TDC, the TKE in the cylinder for D/L = 0.3 was lower than that for D/L = 0.6.

3.3. Effect of D/L on Combustion Characteristics

A difference in the intake port structure leads to changes in the in-cylinder flow field, which will certainly affect the development and propagation of the flame inside the combustion chamber of an RE. Figure 9 illustrates the influence of variations in the D/L on the flame propagation and velocity streamlines in the cylinder at 21°EA BTDC, TDC, 15°EA ATDC, and 48°EA ATDC. It can be seen that at the early stage of combustion, the flame propagation was significantly faster for D/L values of 0.3 and 0.6 than that for D/L = 0.8, and the flame spread area was the largest for D/L = 0.6. This was because the higher TKE in the cylinder for the D/L values of 0.3 and 0.6 promoted flame propagation at the early stage of combustion. After the TDC, the flame propagation area formed at the D/L of 0.6 was the widest, and the spread of the flame to the front and rear end caps partially disappeared at 48°EA ATDC. This showed that the intake port with D/L = 0.6 had the fastest flame propagation velocity and the highest combustion rate. This was attributed to the maximum in-cylinder TKE for D/L = 0.6 during the rapid combustion period. In addition, it could also be seen from Figure 9 that the flame was propagated from the spark plug position to the leading side of the combustion chamber, but it was difficult for it to propagate to the trailing side of the combustion chamber due to the unidirectional flow field, which was the main reason for the high fuel consumption and poor emission of the RE. This finding was also obtained by other researchers [8].

The flame propagation and combustion rate were influenced by the chemical kinetic reaction mechanism of the gasoline during the combustion process, as well as intermediate element reactions, i.e., OH + ${\mathrm{H}}_2$⇌ H + ${\mathrm{H}}_2\mathrm{O}$ and H + ${\mathrm{O}}_2$⇌ O + OH, which directly determined the combustion reaction rate [24]. Figure 10 shows the peak mass fractions of H, O, and OH under different intake port schemes. It can be observed from the figure that when D/L = 0.6, the concentration of active OH radicals produced in the combustion chamber was the highest, which is 53.6% higher than the level with 0.3 and 7.5% higher than the level with 0.8. The increase in OH concentration enabled the concentration of H and O to increase, and this then enhanced the fuel combustion rate. In addition, the heat release of the combustion reaction was mainly generated by the intermediate element reaction of CO + OH ⇌ H + ${\mathrm{CO}}_2$ [25]. Therefore, the increment in the OH production also increased the rate and amount of heat release in the combustion reactions in the combustion chamber, which, in turn, boosted the in-cylinder temperature. This indicated that the intake port with D/L = 0.6 had the maximum heat release and the highest temperature.

The variations in the burned mass fraction in the cylinder according to eccentric angle with different intake port structures are shown in Figure 11. According to the figure, it is easy to notice that the in-cylinder heat release was sequentially delayed for the D/L values of 0.6, 0.8, and 0.3. The intake port with the D/L of 0.6 had the earliest and fastest heat release in the cylinder. The eccentric angle for reaching 90% of the burned mass fraction for the intake port with the D/L of 0.6 was 78.85°EA ATDC, while the eccentric angles for achieving 90% of the burned mass fraction for the D/L values of 0.3 and 0.8 were 120°EA ATDC and 97.98°EA ATDC, respectively. This indicated that the intake port with D/L = 0.6 was the most conducive to promoting chemical reactions and increasing the rate of heat release. This is due to the fact that the intake port with D/L = 0.6 enhanced the local flame propagation rate and shortened the combustion duration.

A comparison of the combustion durations with different intake port structures is shown in Figure 12. EA0, EA50, and EA 90 indicate an eccentric angle corresponding to 0%, 50%, and 90% of the cumulative heat release from fuel combustion in the combustion chamber, respectively. Typically, EA0–10 and EA10–90 denote the flame development and flame propagation of burning mixtures, respectively, and EA50 denotes the combustion efficiency of the fuel. It can be observed from Figure 12 that the values of EA0–10 for the D/L values of 0.3 and 0.6 were significantly smaller than that with the D/L of 0.8, indicating that the flames for the D/L values of 0.3 and 0.6 developed more rapidly at the initial stage of combustion (as shown in Figure 9). Meanwhile, the value of EA10–90 for D/L = 0.6 was decreased by 33.49% and 13.04% in comparison with those for the D/L values of 0.3 and 0.8, respectively, indicating that the intake port with D/L = 0.6 was able to improve the flame propagation speed and promote the combustion reaction in the cylinder. This was mainly due to the fact that the intake port with D/L = 0.6 enabled the vortexes in the cylinder to form earlier and last longer, and the TKE was higher during the combustion stage, which, in turn, promoted mixture formation and flame propagation in the combustion chamber. Furthermore, the value of EA50 for D/L = 0.6 was significantly smaller than those for the D/L values of 0.3 and 0.8, indicating that the in-cylinder combustion’s center of gravity was advanced for D/L = 0.6. This improved the combustion efficiency and reduced combustion losses, which is conducive to more useful work output.

From the in-cylinder mean temperature in Figure 13 and the NO generation curve in Figure 14, it can be seen that the rate and amount of NO generation were consistent with the trend of variation in the in-cylinder mean temperature, which was attributed to the fact that the conditions for NO generation are a high temperature and oxygen-rich atmosphere. Near the TDC, due to the slow flame development and propagation speed (as shown in Figure 9), the temperature in the cylinder changed slowly, and a smaller amount of NO is generated. As the engine operated, the fuels burned rapidly, and the flame propagation speed and heat release increased rapidly, which caused the in-cylinder temperature to rise sharply and drove the rapid generation of NO. For the intake port with D/L = 0.6, the NO generation rate was the fastest, and its peak mass fraction of NO was increased by 197% and 104% in comparison with those for 0.3 and 0.8, respectively. This reason is that, on the one hand, the form of the movement of the mixtures entering the cylinder from the intake port with D/L = 0.6 accelerated the flame propagation (as shown in Figure 9) and improved the fuel combustion rate and heat release rate (as shown in Figure 11), which accelerated the NO generation rate. On the other hand, after the NO generation in the combustion chamber reached the maximum value, it kept a relatively stable value until the exhaust valve was opened.

It can also be concluded from Figure 14 that the amount of NO generated for the intake port with D/L = 0.8 was higher than that for D/L = 0.3. This was mainly due to the fact that when D/L = 0.8, the heat release of fuel combustion was faster (as shown in Figure 11), and the combustion duration was shorter (as shown in Figure 12), which led to a higher peak in-cylinder temperature and promoted the generation of NO. However, this problem can be improved through the post-treatment of exhaust and lean combustion.

4. Conclusions

In this work, a three-dimensional CFD simulation model of a small-scale RE based on a suitable turbulent model and a reduced gasoline chemical reaction mechanism was established. The effects of the taper intake port structure on the flow field distribution and combustion process in the combustion chamber of the small-scale RE were numerically analyzed. The main conclusions that were obtained are as follows:

(1)

The in-cylinder pressure of the RE did not change linearly with the D/L of the intake port in a single direction, but was divided into three levels. The maximum in-cylinder pressure was at the D/L of 0.6, and the minimum in-cylinder pressure was at the D/L of 0.8. In addition, the eccentric angle corresponding to the peak in-cylinder pressure was consistent with the trend of the variation in the peak pressure.

(2)

The gas flow in the intake port of the three representative pressure levels was unidirectional, which caused differences in the vortexes that were formed in the combustion chamber of the RE, and this was mainly reflected in the vortexes that were formed on the leading side of the combustion chamber. This was also the main factor affecting the combustion performance of the small-scale RE. When the D/L was 0.6, the vortexes on the leading side of the combustion chamber formed early and disappeared late, indicating that the vortexes in the combustion chamber lasted longer, and this was conducive to mixture formation.

(3)

The intake port with the D/L of 0.6 promoted the in-cylinder combustion reaction and improved the combustion efficiency of the RE due to the higher flame propagation speed and the earlier center of gravity of combustion, which reduced the combustion losses and facilitated a more useful work output. However, owing to its higher peak in-cylinder temperature, it led to an increase in NO emissions. As for the high levels of emission of NO, these could be reduced through the post-treatment of exhaust and lean combustion.