Scavenge flow analysis of opposed-piston two-stroke engine based on dynamic characteristics

Abstract

Opposed-piston two-stroke engine has been proposed by Beijing Institute of Technology to improve power density and complete machine balance relative to conventional engines. In order to study opposed-piston two-stroke engine scavenging flow, a scavenging system was configured using a three-dimensional computational fluid dynamics model effectively coupled to experiments. The boundary conditions are obtained through one-dimensional working process simulation results and experiments. As the opposed-piston relative dynamic characteristics of opposed-piston two-stroke engine depend on different design and operating parameters including the opposed-piston motion phase difference and crank-connecting rod ratio, a numerical simulation program was built using MATLAB/Simulink to define opposed-piston motion profiles based on equivalent crank angle of opposed crank-connecting rod mechanism. The opposed-piston motion phase difference only affects scavenging timing while crank-connecting rod ratio affects scavenging timing and duration. Scavenging timing and duration are the main factors which affect scavenging performance. The results indicate that a match of opposed-piston motion phase difference and crank-connecting rod ratio has the potential to achieve high scavenging and trapping efficiency with a right flow in cylinder.

Keywords

Uniflow scavenging opposed-piston two-stroke computational fluid dynamics scavenging flow dynamic characteristics

Introduction

Pressured by energy crisis and environmental pollution, car industry is faced with unprecedented challenges due to its high energy consumption and pollution emission.^1,2 In order to keep a sustainable development, car industry will develop in the direction of zero-discharge electric drive vehicle.³ Range-extender electric vehicle is a transitional technology which could overcome the limitation of traditional electric vehicle charging time and improve the travel distance of pure electric vehicles by adding an additional energy storage component.⁴ As a form of range-extender, small generating set is used for online supplement energy, which has a small displacement, high power density, simple, and compact structure. Opposed-piston two-stroke (OP2S) gasoline direct injection (GDI) engine has some advantages in being pure electric vehicle extender, since it is simple, compact, high power density, and good balance.

OP2S engine concept can be traced back to late 1800s. Since then, many novel applications were used on aircrafts, marines, and vehicles. In the first half of the 20th century, OP2S engine was developed in multiple countries for a wide variety of applications. However, modern emission regulations stopped the widespread development of most of the two-stroke engines in the latter half of the 20th century.⁵ In recent years, with the application of advanced design technology, modern analytical tools, materials, and engineering methods, the emission problem is no longer limiting the successful design of a clean and efficient OP2S,⁶ so OP2S engine has again attracted intensive attention to improve engine efficiency and emission performance.^7–11 Compared with conventional engines, OP2S engines have many fundamental advantages.¹² The opposed piston (OP) structure characterized by two pistons reciprocating opposite to each other in a common cylinder canceled the cylinder head and valve mechanism and leads to lower heat loss for higher wall temperature of the two piston crowns compared to a cylinder head. The nearly symmetric movement of OPs leads to excellent engine balance even for single-cylinder configuration.

In recent years, relevant studies were carried out. Stephenson et al. conducted a parametric study on the effects of swirl, initial turbulence, oxygen concentration, and ignition delay on fuel vaporization, mixing, and combustion process. The effects of intake flow field were further investigated by varying the geometry of the intake ports and intake runners in a model dual-port direct injection diesel engine in KIVA simulations.¹³ Stephenson and Rutland¹⁴ assessed the differences in the in-cylinder flow field present during combustion resulting from different intake flow configurations and compared with the significance of spray–wall interaction effects on combustion and emission performance. Bo et al.¹⁵ assessed the accuracy of prediction of the air motion and spray characteristics in a direct injection diesel engine with the computational fluid dynamics (CFD) code. However, gas motion in the OP2S has not been sufficiently studied in the literature. In the OP2S engines, two OPs move symmetrically make the volumetric change rate is twice as large as that of the conventional engines, result in the gas motion and dissipation law are also different. The cylinder head is absence in the OP arrangement, so fuel injector must be installed in the cylinder liner, and the interaction between the fuel spray and the in-cylinder fresh-charge motion with traditionally high swirl resulted in combustion occurring near the combustion chamber surfaces.⁷ Hofbauer conducted a numerical study on the effects of swirl level and nozzle parameters on wall–fuel contact and mixture formation. The results present that low swirl show a favorable effect on the formation.¹⁰ However, since the scavenging process was not included in the calculation, the initial swirl was set as a parameter. To achieve a good swirl level and flow field by designing port structure has not been studied in the literature.

The approach by Beijing Institute of Technology (BIT) uses a uniflow scavenging and GDI technology to achieve separation of injection and scavenging process. The opposed crank-connecting rod mechanism is placed on both sides of cylinder body and a chain transmission mechanism is designed to realize synchronization working of OP. OP2S-GDI engine has some potential advantages such as simple structure, good balance, compact, high power density, and thermal efficiency. The uniflow scavenging system is the most efficient scavenging method for two-stroke engine and is also been considered for free piston engines¹⁶ and OP engines.¹⁷ Numerical modeling of in-cylinder flow in uniflow scavenging system has been the subject of earlier works and include steady-flow simulations,^18,19 two-dimensional (2D)-axisymmetric simulations,²⁰ sector simulations without intake port geometry,²¹ and full three-dimensional (3D) simulations.^22,23 The full 3D simulations are the only approach that allows asymmetric phenomenon to be studied. Multi-dimensional CFD codes are widely used in the design and development of internal combustion engines due to their ability to investigate variables that are difficult or costly to measure in experimental tests.²⁴ CFD offers an expedient means for investigating the flow, gas exchange, and combustion processes under realistic engine operating conditions and identification of important features and major underlying interactions between them.²⁵

Although OP2S engine has already been commonly modeled and tested by many researchers, most of these researches focus on the basic engine performance; they are unable to identify details of the engine design and operating parameters. The article provides some insights into multi-dimensional gas flows in scavenging process of OP2S-GDI engine using commercial CFD software AVL-FIRE based on numerically simulated OP motion profiles and evaluates the scavenging performance using different design and operating parameters, namely, OP motion phase difference and crank-connecting rod ratio. One-dimensional (1D) simulated in-cylinder and scavenge case pressure and temperature were used to define the boundary conditions. The model is validated by comparison of model predictions and experimental data. The mixing of cylinder gas is examined and the concentration of fresh charge in the cylinder at exhaust port closing (EPC) is estimated. The transient behavior of the angular and axial momentum in the engine cylinder is studied over full engine cycle.

Engine modeling

Engine configuration

As shown in Figure 1, OP2S-GDI engine is equipped with GDI system and a uniflow scavenging system, and its injector and spark plug are placed on cylinder liner. On both sides of cylinder liner, there are gas ports, intake ports on one side, and exhaust ports on the other side. Intake ports are used to deliver fresh air into the cylinder, and exhaust ports are used to deliver burnt gas out from the cylinder. In the working process, piston motion controls the opening and closing of ports. There are two pistons placed in cylinder liner, and a combustion chamber is formed when the two pistons move to the closest position. The piston that controls the opening and closing of intake air ports is defined as intake piston and the piston that controls the opening and closing of exhaust air ports is defined as exhaust piston. When the distance between the two pistons is minimized, it is defined as the inner dead center (IDC); when the distance between the two pistons is maximized, it is defined as the outer dead center (ODC). The structure parameters are shown in Table 1.

Figure 1.

OP2S-GDI engine.

Table 1.

Engine specifications.

Parameters	Sym.	Unit	Value
Bore	B	mm	56
Stroke	S	mm	49.5 (×2)
Connecting rod	L	mm	82.5
Effective compression ratio	CR	–	10.5
Engine speed	N	r/min	6000
Number of intake ports	n_in,port	–	10
Number of exhaust ports	n_ex,port	–	10
Intake port height	h_in	mm	12
Exhaust port height	h_ex	mm	14
Intake port circumference ratio	β_in	–	0.75
Exhaust port circumference ratio	β_ex	–	0.6
Intake port radial angle	$ϕ_{in}$	°	15
Exhaust port radial angle	$ϕ_{ex}$	°	0
Power	P	kW	15
Fuel consumption rate	BSFC	g/kW h	276

BSFC: brake-specific fuel consumption.

Scavenging description

Figure 2 illustrates the principle of uniflow scavenging process of OP2S gas exchange. The exhaust port opens first before ODC and a blow down discharge process commences. The discharge period up to the time of the scavenging port opening is called the free exhaust period. The intake port opens also before ODC when the cylinder pressure slightly exceeds scavenging pressure. Because the burned gas flow is toward the exhaust port, which is now having a large open area, the exhaust flow continues and few backflow occurs. When the cylinder pressure is less than the scavenging pressure, fresh air enters the cylinder and the scavenging process starts. This flow continues as long as the intake port is opened, and the scavenging total pressure exceeds the pressure in the cylinder. As the pressure rises above the exhaust pressure, the fresh charge flowing into the cylinder displaces the burned gas: a part of the fresh air mixes with the burned gas and is expelled with them. The intake port closes after the exhaust port closes; since the flow is toward the intake port continuously, additional fresh air is obtained. The additional air inflow period up to the time of intake port close is called the post intake period.

Figure 2.

Gas exchange in OP2S engine.

OP2S-GDI engine dynamic

In the application of the current mainstream CFD code such as star-CD and AVL-FIRE, researchers need structural parameters of crank-connecting rod system or a predefined piston motion profile to simulate piston movement. As the piston relative motion profile of OP2S engine is different from that of conventional engine, the existing work on the CFD modeling of OP2S engines uses piston motion profiles obtained from a dynamic engine model. In this article, piston motion profile was obtained from a dynamic calculation which is built by MATLAB/Simulink. As shown in Figure 3, due to OP motion phase difference, OPs on both sides cannot arrive at each top dead center (TDC) simultaneously. If the phase difference of the intake and exhaust pistons is $ϕ$ , the relative displacement of OP should be known by the kinematics of traditional crank-connecting rod mechanism

\begin{matrix} X & = X_{in} + X_{ex} \\ = r (1 - \cos (α - \frac{φ}{2}) + \frac{1}{2} λ_{in} \sin^{2} (α - \frac{φ}{2})) \\ + r (1 - \cos (α + \frac{φ}{2}) + \frac{1}{2} λ_{ex} \sin^{2} (α + \frac{φ}{2})) \end{matrix}

(1)

where $X_{in}$ and $X_{ex}$ are the displacement of the intake and exhaust pistons, respectively; $r$ is the crank radius; $λ = r / l$ is the crank-connecting rod ratio; $l$ is the length of the connecting rod; and $α$ is the equivalent crank angle (CA).

Figure 3.

Opposed crank-connecting rod mechanism.

When $α$ is 0°CA, the distance between the two pistons is minimized, which is defined as the IDC. When $α$ is 180°CA, the distance between the two pistons is maximized, which is defined as the ODC.

Taking the derivative of equation (1) about $α$ , the relative velocity of OP can be obtained as follows

\begin{matrix} X' = & X'_{in} + X'_{ex} = r (\sin (α - \frac{φ}{2}) + \frac{λ_{in}}{2} \sin 2 (α - \frac{φ}{2})) \\ + r (\sin (α + \frac{φ}{2}) + \frac{λ_{ex}}{2} \sin 2 (α + \frac{φ}{2})) \end{matrix}

(2)

Taking the derivative of equation (2) about, $α$ the relative acceleration of OP can be obtained as follows

\begin{matrix} X ″ = & X ″_{in} + X ″_{ex} = r (\cos (α - \frac{φ}{2}) + λ_{in} \cos 2 (α - \frac{φ}{2})) \\ + r (\cos (α + \frac{φ}{2}) + λ_{ex} \cos 2 (α + \frac{φ}{2})) \end{matrix}

(3)

CFD model and setup

AVL-FIRE software is used to build CFD model in working process simulation. The dynamic mesh tool Fame Engine is used to generate the moving meshes of the cylinder by defining moving selection, buffer selection, interpolation selection, and the relative motion rule of OP. Intake and exhaust chambers generated no moving meshes which are refined near the intake and exhaust ports, in order to capture the significant flow gradients accurately. The moving parts were meshed with layered hexahedrons with the layers being normal to the movement of the piston. Since the geometry of the scavenging port was very irregular, it was meshed with hybrid grids, including tetrahedrons, prisms, and hexahedrons to avoid divergence and ensure the accuracy of the results. Figure 4 shows the full-scale 3D CFD model consisting of 249,528 cells at scavenging process and 47,961 cells at compression process after rezone. Dynamic mesh of piston motion in the intake and exhaust strokes has been treated according to the realistic motion rule of OP. The scavenging calculation is from exhaust port opening (EPO) to intake port closing (IPC), while the in-cylinder process is from IPC to EPO.

Figure 4.

CFD calculation model.

Mesh movement includes three parts: left and right pistons to simulate the gas motion during the entire working process. $k - ε$ model was used in the calculation of turbulence. The in-cylinder pressure, velocity, temperature, and concentration fields are predicted through the solution of the Reynolds-averaged Navier–Stokes (RANS) equations. The time step for the calculation is set to about 0.1°CA–0.4°CA for all the calculation cases depending on convergence at each time step. Discretization was achieved through the second-order MINMOD relaxed differencing scheme for momentum and continuity, while the first-order upwind differencing scheme was used for turbulence, energy, and scalar quantities. During a simulation, the CFD solver needed to know when to jump to the next time step for a transient run. There were two ways the solver could make this decision. The pressure and momentum were solved until the reduction of residuals reached 10⁻⁴ or the total number of iterations exceeded 60; the solver would then jump to the next time step.

Boundary and initial conditions

The boundary conditions were chosen to reflect the physical conditions in the validation model and the prototype engine. A constant pressure boundary condition is used for both intake and exhaust ports. The mean scavenge pressure is taken as 1.2 bar and the mean exhaust receiver pressure is taken as 1 bar. Frictional effects at the walls are not taken into account, that is, the smooth wall option for turbulent flow boundary condition is used. The initial conditions in the cylinder are extracted from the GT-Power software simulation for every case. The flow field is initialized by specifying the temperature, pressure, and turbulence intensity. By performance prediction, initial in-cylinder pressure and temperature are computed in the case of 15 kW at 6000 r/min, which are the initial conditions for CFD. The initial temperatures of cylinder, intake chamber, and exhaust chamber are given a value of 784 K, 322 K, and 634 K, respectively. The constant wall temperatures were also used.

Model validation

Uniflow scavenging process is assumed to be completed in three models: perfect scavenging, perfect mixing, and short circuit. In practice, scavenging process includes multiple scavenging models, giving a relation for scavenging efficiency

η_{S} = {\begin{matrix} l_{0} & l_{0} \leq l_{c} \\ 1 - e^{- i \cdot l_{0}} & l_{0} > l_{c} \end{matrix}

(4)

where $η_{S}$ is the scavenging efficiency, $l_{0}$ is the delivery ratio, $i$ is the scavenging model index, $l_{c}$ is the demarcation point between perfect scavenging and rich exhaust scavenging, and e is the power exponent.

The boundary and initial conditions of CFD calculation are given by 1D simulation. Meanwhile, coupling simulation characteristic of 1D and 3D can be adopted for model testing.⁹ The experiment carried out 1D and 3D scavenging process comparison validation on the above five cases at 6000 r/min. As shown in Table 2, given scavenging model index $i$ , delivery ratio and scavenging efficiency derived from 1D performance prediction and 3D numerical simulation show a good agreement.

Table 2.

Comparison between scavenging results.

Case	$φ$	$λ$	$i$	Delivery ratio		Scavenging efficiency
Case	$φ$	$λ$	$i$	1D	3D	1D	3D
1	5	0.28	2.21	1.08	0.994	0.704	0.719
2	15	0.28	1.44	1.21	1.196	0.511	0.528
3	25	0.28	1.91	1.50	1.467	0.866	0.857
4	15	0.24	2.93	0.99	1.050	0.882	0.878
5	15	0.32	2.67	1.10	1.075	0.874	0.886

1D: one-dimensional; 3D: three-dimensional.

When $φ$ is 15°CA and $λ$ is 0.28, a simulation is performed at 20% load of 1200 r/min. A series of comparison validation in 1D in-cylinder working process, 3D scavenging process, and motored conditions was conducted, as shown in Figure 5. The simulation results agree with the experiment results in scavenging process which indicate that the parameters are reasonably selected.

Figure 5.

Comparison between in-cylinder pressures.

Simulation results and discussion

Scavenging process

Piston relative motion profile of OP2S is determined by $φ$ and $λ$ . The effect of $φ$ and $λ$ on the piston displacement is shown in Figure 6. The variation in the scavenging timing and duration over the engine cycle is presented in Table 3. The $φ$ only influences the scavenging timing and the scavenging duration is constant. The increase in $φ$ can enlarge the interval of EPO and intake port opening (IPO) and IPC; meanwhile, the opening and closing of exhaust port are advance and the opening and closing of intake port are delay. The free exhaust period and the post intake period increase. The $λ$ influences the scavenging timing and duration. With the increase in $λ$ , the opening and closing of exhaust port are advance and the opening and closing of intake port are delay. The free exhaust period and the post intake period can be constant, but the scavenging duration increases.

Figure 6.

Intake and exhaust piston displacement: (a) effect of $φ$ on piston displacement and (b) effect of $λ$ on piston displacement.

Table 3.

Variation in the scavenging timing and duration.

Case	$φ$	$λ$	Scavenging timing				Scavenging duration		Free exhaust period	Post intake period
Case	$φ$	$λ$	EPO	IPO	EPC	IPC	Exhaust	Intake	Free exhaust period	Post intake period
1	5	0.28	102	112.5	253	252.5	151	140	10.5	−0.5
2	15	0.28	97	117.5	248	257.5	151	140	20.5	9.5
3	25	0.28	92	122.5	243	262.5	151	140	30.5	19.5
4	15	0.24	101.5	122	243.5	253	142	131	20.5	9.5
5	15	0.32	92.5	113	252.5	262	160	149	20.5	9.5

EPO: exhaust port opening; IPO: intake port opening; EPC: exhaust port closing; IPC: intake port closing.

The mass flow through intake ports during scavenging is presented in Figure 7. The mass flow increases rapidly at first after IPO and reaches a maximum with slow fluctuations before ODC. After mass flow reaches a maximum, the mass flow shows a linear decreasing trend. In-cylinder pressure at IPO depends on the time of EPO and the free exhaust period and the in-cylinder pressure at IPC depends on the time of EPC and the post intake period. A longer EPO advance and a shorter intake closing delay can avoid the blow black of burnt gas and fresh charge at the time of IPO and closing. A larger $ϕ$ enables a longer exhaust opening advance and intake closing delay which makes a lower flow resistance in the beginning of the scavenging process and a higher flow resistance in post intake period. As shown in Figure 7(a), when $φ$ is 5°CA, the negative mass flow rate is observed after IPO shows that blow black occurs; when $ϕ$ is 25°CA, the negative mass flow rate is observed after EPC shows that blow black occurs. A higher $λ$ enables a longer scavenging duration and an early intake and exhaust opening. In addition, a higher $λ$ increases backflow at the beginning of scavenging process and post intake period, as shown in Figure 7(b). When $λ$ is 0.32, the negative mass flow rate is observed after IPO and EPC shows that blow black occurs.

Figure 7.

Mass flow through the intake port: (a) effect of $φ$ on intake mass flow and (b) effect of $λ$ on intake mass flow.

Figure 8(a) shows the $φ$ on the trapping and scavenging efficiency when the engine speed is 6000 r/min. The scavenging efficiency increases rapidly with $φ$ increasing from 5°CA to 15°CA and the scavenging efficiency increases slowly while $φ$ varies from 15°CA to 25°CA. However, the trapping efficiency increases while $φ$ varies from 5°CA to 15°CA and drops while $φ$ varies from 15°CA to 25°CA and is maximized as $φ$ is 15°CA. When the intake port is opening, the lower the cylinder pressure is, the more beneficial to intake it is. But when the cylinder pressure is low to a certain extent, it is bad for intake trapping. So the trapping efficiency begins to decrease when $φ$ is larger than 15°CA. Meanwhile, with the increase in intake port delay angle, the delivery ratio and scavenging efficiency are increased accordingly. Figure 8(b) shows the effect of $λ$ on the trapping and scavenging efficiency when the engine speed is 6000 r/min. The scavenging efficiency increases gradually with $λ$ increasing from 0.24 to 0.32 which is approximate linear characteristics. However, the trapping efficiency increases while $λ$ varies from 0.24 to 0.28 and drops while $λ$ varies from 0.28 to 0.32 and is maximized as $λ$ is 0.28. When the intake port is opening, the lower the cylinder pressure is, the more beneficial to intake it is. A higher $λ$ results in an earlier open and later close of the intake and exhaust ports and increases scavenging process. When scavenging process is too long, the loss of the amount of the fresh charge is increased and the trapping efficiency is decreased.

Figure 8.

Scavenging results of different design and operating parameters: (a) effect of $φ$ on scavenging process and (b) effect of $λ$ on scavenging process.

Scavenging flow

Figure 9(a) shows when OP is around IDC, the greater the $φ$ is, the smaller the relative velocity is. This is caused by $φ$ , which results in the exhaust piston moving away from IDC, and the intake piston will close to IDC. There is a catch-up process between the two pistons, and when the bigger the $φ$ is, the longer the catch-up process will be. A larger $φ$ is beneficial to the constant volume combustion and thermal efficiency. Figure 9(b) shows when the OP is around IDC, the greater the $λ$ is, the greater the relative velocity is. However, when the OP is nearby ODC, the greater the $λ$ is, the smaller the relative velocity is. Therefore, the larger the $λ$ is, the relative velocity of piston around ODC is smaller, which is beneficial to the scavenging efficiency but not conducive to trapping efficiency.

Figure 9.

Intake and exhaust piston relative velocity: (a) effect of $φ$ on piston relative velocity and (b) effect of $λ$ on piston relative velocity.

This special dynamic characteristic may influence OP2S engine performance by affecting gas motion in the cylinder. As shown in Figure 10, due to relative movement of OPs in compression and expansion process, in-cylinder gas momentum near cylinder middle section is counteracting. Affected by phase difference, piston movement velocity is not 0 and the direction is the same around IDC during intake and exhaust. Therefore, air axial motion velocity is not 0. As shown in Figure 10(a), $φ$ affects the relative movement of OPs and minimum working volume around IDC. The larger the $φ$ is, the longer it takes in OP concurrent catch-up and the higher the velocity is. Therefore, in-cylinder gas axial motion velocity is bigger around IDC. As shown in Figure 10(b), different $λ$ only affects the relative movement of piston rather than minimum working volume. Consequently, when $λ$ increases, absolute catch-up velocity increases as OP concurrent catch-up continues. Therefore, it can be concluded that in-cylinder gas axial motion velocity increases around IDC.

Figure 10.

Axial velocity of gas in the engine cylinder: (a) effect of $φ$ on axial velocity and (b) effect of $λ$ on axial velocity.

Figure 11 shows the effect of $φ$ and $λ$ on in-cylinder swirl ratio. Different $φ$ affects scavenging timing and in-cylinder thermodynamic state as well during intake phase. As shown in Figure 11(a), a larger $φ$ can advance exhaust and prolong free exhaust process so that in-cylinder pressure is lower and intake velocity increases during intake phase. As a result, in-cylinder swirl ratio increases rapidly. If a higher in-cylinder pressure is formed in intake process, burnt gas backflow is likely to occur. During compression process after intake port closes, swirl ratio decreases gradually and tends to be uniform because in-cylinder swirl is compressed and sheared. The results show that $φ$ only affects the crankshaft position of maximum swirl ratio while little on swirl ratio value. The $λ$ affects scavenging timing, duration, and thermodynamic state as well during intake phase. As shown in Figure 11(b), a higher $λ$ prolongs intake and exhaust close, which results in a continuous rising momentum in intake where in-cylinder swirl intensifies reaching its maximum at 40°CA after ODC. During compression process, swirl ratio decreases linearly. The result shows that $λ$ only affects in-cylinder swirl ratio rather than the crankshaft position of maximum swirl ratio.

Figure 11.

Swirl ratio of gas in the engine cylinder: (a) effect of $φ$ on swirl ratio and (b) effect of $λ$ on swirl ratio.

Turbulence kinetic energy (TKE) is defined as mean kinetic energy per unit mass associated with eddies in turbulent flow. Figure 12 shows that the in-cylinder mean TKE changes tend to be uniform in compression process. The result can be attributed to two reasons: as cylinder axial motion velocity increases, small vortex breaks into turbulence; since the velocity of piston motion is nonzero in IDC and gas exchange near pistons is active, TKE increases rapidly. When piston motion moves to IDC, mean TKE changes slowly due to eddy shearing and dissipation. The result also finds that maximum mean TKE grows with the in increase in OP relative velocity before IDC. At the IDC, although minimum working volumes of different $φ$ , mean TKE tends to be uniform quantitatively as shown in Figure 12(a); minimum working volumes at different $λ$ remain the same and TKE decrease resembles, as shown in Figure 12(b).

Figure 12.

TKE of gas in the engine cylinder: (a) effect of $φ$ on TKE and (b) effect of $λ$ on TKE.

Angular and axial gas motion

To quantify the transient behavior of the system, the volume integrated angular and axial momentum in the engine cylinder are evaluated when $φ$ is 15°CA and $λ$ is 0.28. The angular momentum L_z and axial momentum G are given by

L_{z} = \int_{V} r ρ v_{θ} dV

(5)

G = \int_{V} ρ v_{z} dV

(6)

where V is the instantaneous cylinder volume, $v_{θ}$ is the angular velocity of volume element, r is the turning radius of volume element, $v_{z}$ is the axial velocity of volume element, and $ρ$ is the gas density.

The variation in the angular and axial momentum over the engine cycle is presented in Figure 13 using the angular momentum at ODC L_z , _ODC and the cylinder radius B/2 for normalization. As the exhaust starts at EPO, a rapid decrease in the angular momentum is observed in Figure 13(a). The decrease is caused by burnt gases being blown out through the exhaust port due to the high in-cylinder pressure. As the fresh charge is blown into the cylinder with a tangential velocity component, a rapid increase in the angular momentum is observed. The maximum angular momentum is obtained approximately after ODC. Before EPC, an unexpected reduction in the angular momentum is observed. This drop is mainly caused by the decrease in intake mass flow velocity due to the increase in in-cylinder pressure and decrease in gas ports flow area by the inward motion of the OP. From IPO to ODC, the change in angular momentum shows a delay caused by the flow inertial. The angular momentum is minimized at 8°CA after IPO and the angular momentum is maximized at 40°CA after ODC. Figure 13(b) shows that the axial momentum of the cylinder charge is not 0 at IDC because the absolute velocity of OP is not 0 unlike the conventional engine at TDC when the piston motion phase difference is not 0. As the piston begins to move outward, the fluid gain momentum is reduced. At EPO, the cylinder charge is directed out through the exhaust port and the axial momentum increases rapidly. At IPO, the cylinder charge again gains momentum due to the fresh charge entering from the intake port. After ODC, the axial momentum decreases as the mass flow rate through the intake ports decreases. The closing of the exhaust port creates pressure waves that travel between the OP surfaces. The pressure waves are observed as small superimposed oscillations on the curve after EPC and IPC. At IPC, a pressure wave is again created traveling between the OP surfaces. A decrease at 30°CA after ODC is, however, observed, corresponding to a decrease in mass flow rate. After EPC, the oscillations decrease and the pressure increases in the compression process.

Figure 13.

Angular and axial momentum in the cylinder: (a) angular momentum in the engine cylinder and (b) axial momentum in the engine cylinder.

Conclusion

In this article, a CFD model is developed, validated, and applied for investigating the scavenge flow in the OP2S uniflow scavenging engine. The model gives detailed information on the temporal development of in-cylinder flow field with different design and operating parameters, namely, $φ$ and $λ$ . The main findings of the work are listed below.

The $φ$ only affects scavenging timing rather than duration. Smaller $φ$ shortens free exhaust process leading to a possible burnt gas backflow after IPO while overlarge $φ$ prolongs post intake period producing fresh charge backflows before IPC which reduce trapping rate and scavenging efficiency.

Different $λ$ affects scavenging timing and duration. When $λ$ exceeds significantly, burnt gas and fresh charge backflow after IPO and before IPC, respectively, thus reducing trapping rate. In contrast, when $λ$ is too small, scavenging efficiency is decreased due to a shortened scavenging duration.

The $ϕ$ affects the relative motion of piston and its minimum working volumes and $λ$ affects the relative motion of piston rather than minimum working volumes. The larger $ϕ$ is, the longer it takes in OP catch-up and the higher the air axial velocity is around IDC. With the increase in $λ$ , the OP catch-up velocity and gas axial velocity increase around IDC.

During compression process, swirl ratio decreases gradually due to wall friction and dissipation. The $ϕ$ only affects crankshaft position of maximum swirl ratio and $λ$ only affects the value of maximum swirl ratio.

The angular momentum of in-cylinder gas is delayed by 8°CA after IPO due to inertia effect on in-cylinder flow and the maximum angular momentum occurs at 40°CA after ODC.

Footnotes

Academic Editor: Jose Ramon Serrano

Declaration of conflicting interests

The authors declare that they have no conflict of interests regarding the publication of this article.

Funding

This research received no specific grant from any funding agency in the public, commercial, or not-for-profit sectors.

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McGough

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Wahl

Regner

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10.

Regner

Herold

Wahl

. The achates power opposed-piston two-stroke engine: performance and emissions results in a medium-duty application. SAE technical paper 2011-01-2221, 2011.

11.

Redon

Kalebjian

Kessler

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12.

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13.

Stepheson

Claybaker

Rutland

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14.

Stephenson

Rutland

. Modeling the effects of valve lift profile on intake flow and emissions behavior in a direct injection diesel engine. SAE technical paper 952430, 1995.

15.

Clerides

Gosman

. Prediction of flow and spray processes in an automobile direct injection diesel engine. SAE technical paper 970882, 1997.

16.

Mikalsen

Roskilly

. A computational study of free-piston diesel engine combustion. Appl Energ 2009; 86: 1136–1143.

17.

Wang

Zhu

. Numerical analysis of two-stroke free piston engine operating on HCCI combustion. Appl Energ 2011; 88: 3712–3725.

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Hossain

Zhang

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19.

Tamamidis

Assanis

. Optimization of inlet port design in a uniflow scavenged engine using a 3-D turbulent flow code. SAE technical paper 931181, 1993.

20.

Ravi

Marathe

. Effect of inlet and exhaust pressures on the scavenging characteristics of a carbureted uniflow scavenged engine. SAE technical paper 920840, 1992.

21.

Diwakar

. Three-dimensional modeling of the in-cylinder gas exchange processes in a uniflow-scavenged two-stroke engine. SAE technical paper 870596, 1987.

22.

Goldsborough

Van Blarigan

. Optimizing the scavenging system for a two stroke cycle, free piston engine for high efficiency and low emissions: a computational approach. SAE technical paper 2003-01-0001, 2003.

23.

Lamas

Rodríguez Vidal

. Computational fluid dynamics analysis of the scavenging process in the MAN B&W 7S50MC two-stroke marine diesel engine. J Ship Res 2012; 56: 154–161.

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