Matching analysis of combined charging of an opposed-piston two-stroke engine

DONG Xue-fei(董雪飞), ZHAO Chang-lu(赵长禄), ZHANG Fu-jun(张付军), ZHAO Zhen-feng(赵振峰), WU Tao-tao(吴滔滔)

(Fundamental Science on Vehicular Power System Laboratory, Beijing Institute of Technology, Beijing 100081, China)



Matching analysis of combined charging of an opposed-piston two-stroke engine

DONG Xue-fei(董雪飞), ZHAO Chang-lu(赵长禄), ZHANG Fu-jun(张付军), ZHAO Zhen-feng(赵振峰), WU Tao-tao(吴滔滔)

(Fundamental Science on Vehicular Power System Laboratory, Beijing Institute of Technology, Beijing 100081, China)

To experimentally match performance and structural features of an opposed-piston two-stroke engine (OPTSE), two calculation models, a one-dimensional (1-D) model and a three-dimensional (3-D) model, of the combined charging matching simulation of an OPTSE was established by using the GT-Power software. To test and verify the one dimensional model, the three-dimensional computational fluid dynamics simulation model was established using AVL FIRE software. Cylinder pressure curves in these two models match perfectly, showing that it is reasonable to use the one-dimensional model to simulate the work process of an OPTSE. Moreover, the effects of delivery ratio, nozzle ring diameter and exhaust back pressure on brake specific fuel consumption (BSFC) were studied.

opposed-piston; combined charging; uniflow scavenging; delivery ratio； exhaust back pressure

Unlike four-stroke engines, the exhaust gas in an opposed-piston two-stroke diesel engine (OPTSE) is not pushed out from the cylinder, but swept by the pre-compressed air. So the air must be provided by an external scavenge blower[1]which consumes some of the engines power, though turbochargers can perform this function at some operating points. The blower power requirement reduces the engine’s brake output and increases brake specific fuel consumption (BSFC) of the engine.

Given the scavenging characteristics of an OPTSE, principle prototype experiments adopt combined charging to simulate the intake, exhaust, and charging systems. In this paper, the combined charging matching and the influence of the major parameter on BSFC were studied.

Fig.1 Principle construction diagram of an OPSDE

1 OPTSE principle and construction

An OPTSE has three features: uniflow scavenging, asymmetric port timing, and controlled boosting. The uniflow scavenging is achieved with the intake and exhaust ports placed on the two extreme sides of the cylinder. The principle construction diagram of an OPTSE is shown in Fig.1 and the characteristics are given in Tab.1.

For OPTSE without crankcase scavenging, assistance to provide the intake boost pressure is necessary to start the engine before the turbine receives enough energy from the exhaust gas to drive the compressor. The engine is thus supercharged and turbo compounded rather than turbocharged in the true sense.

Tab.1 Engine characteristics

2 Supercharging system

The basic design of the gas exchange system (Fig.2) consists of a variable geometry turbocharger (VGT). A Rotrex centrifugal compressor is geared to the motor and controlled by electronic control unit (ECU) and an intercooler. The Rotrex centrifugal compressor also reduces intake system weight as well as the duration for the assisted boosting during the engine startup period. At a transient state when the engine needs to increase torque output, compared to the blower geared to the crankshaft, the centrifugal compressor driven by the motor would increase the intake pressure much faster and result in a faster engine transient response.

Fig.2 Layout scheme of the gas exchange system

3 Modeling and matching calculation

3.1 Modeling equivalent and assumption

Given that there is no physical model available for OPTSE, an equivalent transformation from traditional engine model was proposed in this paper, based on the following three principles: ①the piston characteristic of motion is invariant; ②the cylinder geometry parameters are unchangeable, including cylinder bore, stroke, and compression ratio; ③working volume change characteristic remains the same as the OPTSE working volume, this is because that the volume change law would affect the calculation precision of in-cylinder pressure. The equivalent transformations were proceeded as follows.

① Two opposed pistons in the cylinder were replaced by only one piston whose to crankshaft angle were now the sum of the previous two pistons.

② The TDC are defined as the point when the relative displacement between two opposed pistons is 0. By setting this point as a benchmark, the piston relative displacements in the cycle could be determined. More details can be seen in Fig.3.

Fig.3 Equivalent model transformation

③ Suppose that working medium state in cylinder is homogenous, working medium is perfect gas, the gas that inflows and outflows of cylinder is 1-D compressible flow and the kinetic energy of working medium at the inlet port and outlet port is negligible. Inside the system boundary, the thermodynamic state and chemical composition at different positions were considered to be exactly the same, namely, 0-D model. Atmospheric temperature and pressure are defined according to international standards.

3.2 1-D modeling and setting boundary

The 1-D GT-POWER simulation model is established[2], with the model settings and the boundary conditions shown in Tab.2.

Tab.2 Main initial parameters

3.3 Model verify

To test and verify the one-dimensional model, the three-dimensional computational fluid dynamics simulation model was established using AVL FIRE software[3].

During the gas exchange process, air delivered is complex three-dimensional viscous flow, and there are vortex, separation and other complex flow phenomenon[4]. So thek-ζ-fmodel is chosen as the turbulence model, which has been developed by Hanjalic, Popovac and Hadziabdic[5]. And the Eddy Break-up model as described by Magnussen and Hjertager[6]is chosen as the combustion model. The chemical reactions usually have time scales that are very short compared to the characteristics of the turbulent transport processes. Thus, it can be assumed that the rate of combustion is determined by the rate of

intermixing on a molecular scale of the eddies containing reactants and those containing hot products.

The Meshes model based on AVL FIRE is established as shown in Fig.4, and the boundary conditions are shown in Tab.3.

Fig.4 Meshes of the scavenging system

Tab.3 Boundary conditions of the CFD mode

According to Tabs.1-3, and Refs.[7-8], the initial parameters are setup, and the two kinds of simulation software define the same boundary conditions for the simulation calculation.

Fig.5 shows the comparison of 1-D GT-Power and 3-D CFD simulation pressure curves in cylinder for a selected engine operation condition, and the cylinder pressure curves in these two models match perfectly. So it is reasonable to use the one-dimensional model to simulate the work process of OPTSE. With high confidence gained, the calibrated GT-Power simulation model is not only used for future predictions but is also very helpful in identifying any weak points and, therefore, directions for future improvements.

Fig.5 Comparison between 1-D GT-Power and 3-D CFD simulation results

4 Simulation results

4.1 Parameter study

The purpose of this simulation study is to seek the optimum match between the opposed-piston two-stroke engine and supercharging system. Parameter studies are proceeded under three engine speeds, so as to consider the influences of delivery ratio (Δ), turbine orifice diameter (Dt) and exhaust back pressure (Pex) on brake specific fuel consumption and trapped A/F. In this study, under the same speed, no matter how these parameters changed, the output power should be ensured to be the same. Tab.4 shows the three parameters range and corresponding working conditions.

Tab.4 Simulation parameters input

The centrifugal compressor is not geared to the crankshaft but driven bya motor, and a key evaluation indicator in the OPTSE performance matching is BSFC (b) which can be calculated from engine speed (n), fuel consumption (m) per cycle, brake effective power (P) and scavenging blower consumed power (Pe).

(1)

4.1.1 Effect of turbine orifice diameter on BSFC and trapped A/F

The influence of turbine orifice diameter on BSFC and trapped A/F is shown in Fig.6 and Fig.7.

Fig.6 Influence of turbine orifice diameter on BSFC

Fig.7 Influence of turbine orifice diameter on trapped A/F

The simulation results show that turbine orifice diameter has a strong influence on the engine BSFC and a smaller influence on trapped A/F in these three typical cases, so the fuel economy and the trapped A/F could be improved via increasing the turbine orifice diameter as shown in Fig.7 and Fig.8.

Fig.8 Influence of exhaust back pressure on BSFC

When increasing the turbine orifice diameter with constant delivery ratio and exhaust manifold pressure, the flow mass and pressure of compressed gas increase, then Rotrex supercharger consumed power reduces, which results in BSFC decrease.

However, the diameter range is limited by exhaust pressure and the trapped A/F which must achieve the designed target in the entire engine speed range. With the diameter increasing, the turbine flow capacity is amplified, thus maintaining the high exhaust pressure is difficult at lower engine speed. For example, in order to keep thePex=0.145 MPa in the scavenging phase and the trapped A/F>18 at 1 600 r/min, the optimum diameter range is from 33 mm to 39 mm.

4.1.2 Effect of exhaust back pressure on BSFC and trapped A/F

The influences of exhaust back pressure on BSFC and trapped A/F are shown in Fig.8, and Fig.9.

Fig.9 Influence of exhaust back pressure on trapped A/F

The simulation results show that exhaust back pressure has a strong influence on trapped A/F and a smaller influence on BSFC in the three typical cases, so the main method of increasing the trapped A/F is to improve the exhaust pressure. When increasing exhaust pressure with the constant delivery ratio and turbine orifice diameter in the three typical cases, the pressure in the cylinder at the end of the scavenging process increase and the density rises, and then the trapped A/F increases.

When decreasing exhaust pressure(<0.16 MPa) at 2 500 r/min in Fig. 9, the pressure in the cylinder at the end of the scavenging process decrease, which decreases the trapped air mass and worsens in-cylinder combustion. To output the same power from crankshaft, the consumed fuel mass increases, then the BSFC rise. However, with exhaust pressure(>0.16 MPa) continuing to go up at 2 500 r/min, the pressure difference of the inlet port and outlet port must be increased to maintain the constant delivery ratio, then super charger consumed power increases, which results in BSFC rise.

4.1.3 Effect of delivery ratio on BSFC and trapped A/F

Fig.10 Influence of delivery ratio on BSFC

Fig.11 Influence of delivery ratio on trapped A/F

The simulation results of delivery ratio on BSFC and trapped A/F are shown in Fig.10 and Fig.11. The simulation results show that delivery ratio has a strong influence on BSFC and little influence on trapped A/F in the three typical cases. At the different engine speeds the influence is that the higher the speed, the larger the influence on BFSC. The reason is that, with increasing engine speed, scavenging time interval per cycle reduces continuously and the air consumption improves continually. Therefore the pressure difference between inlet port and outlet port needs to be increased to maintain the constant delivery ratio, resulting in BSFC increase.

Fig.12 Delivery ratio in the entire engine speeds

Delivery ratio is limited by inlet manifold pressure and exhaust manifold pressure. With the constant exhaust manifold pressure, inlet manifold pressure has a stronger influence on delivery ratio. The larger the scavenging pressure is, the larger the delivery ratio is, which results in the BSFC rises. To the contrary, if the delivery ratio is too small, the exhaust gas is unable to be cleaned out the cylinder completely. So it is important to ensure the appropriate delivery ratio at different engine speeds. The delivery ratios are confirmed by Ref.[9] as shown in Fig.12.

4.2 Engine performance prediction

With increasing engine speed, the consumed power by Rotrex supercharger rises as shown in Fig.13. The main reason is that the higher engine speed is, the shorter space of time of the scavenging process per cycle is. So it is necessary to improve the pressure difference between inlet port and outlet port by using a supercharger, especially at high engine speeds as shown in Fig.13. Because the influence of Rotrex supercharger is larger at high speed ranges, Rotrex is used in the entire engine speed ranges.

Fig.13 Simulated full load torque curve, Rotrex power curve and differential pressure curve

Fig.14 Flow matching characteristic of supercharger and engine at full load

Compressor working points are plotted on the manufacturer-provided compressor map as shown in Fig.14 and Fig.15, which demonstrate a very good match between the engine and the compressor map.

Fig.15 Flow matching characteristic of turbocharger and engine at full load

5 Conclusions

Based on the simulation results, the following conclusions can be made:

① The combined charging system is effective to improve engine performance and satisfies the corresponding delivery ratio demand at the entire engine speed ranges.

② Compared with exhaust back pressure, theturbine orifice diameter and delivery ratio have the major influence on BSFC. The main way to decrease BSFC is to increase the turbine orifice diameter in a certain range or reduce the delivery ratio.

③ Compare with delivery ratio and the turbine orifice diameter, the exhaust back pressure has larger influence on trapped A/F. The main way to produce more power is to improve the exhaust back pressure appropriately.

[1] Jean P, Martin F. Opposed piston engines: evolution, use, and future applications[M]. Warrendale:

SAE International, 2010: 17-80.

[2] Liu Yongchang. A modification to the “fully mixed/layers formed” compound scavenging model in a two-stroke diesel engine with constant pressure Charging[J]. Internal Combustion Engine Engineering, 1985, 4(2): 33-41. (in Chinese)

[3] Zhu Tao, Wang Yang, Zhang Zhong. Gas flow simulation of hydraulic free piston engine[J]. China Mechanical Engineering, 2010, 21(18):2196-2201. (in Chinese)

[4] Chen Wenting, Zhu Weilin, Zhang Yangjun， et al. Simulation on the scavenging process of the opposite two-stroke diesel engine[J]. Journal of Aerospace Power, 2010, 25(6): 1322-1326. (in Chinese)

[5] Hanjalic K, Popovac M, Hadziabdic M. A robust near-wall elliptic-relaxation eddy-viscosity turbulence model for CFD [J]. International Journal of Heat and Fluid Flow, 2004,25(6):1047-1051.

[6] Magnussen B F, Hjertager B H. On mathematical modeling of turbulent combustion with special emphasis on soot formation and combustion[C]∥Sixteenth International Symposium on Combustion, Pittsburgh, 1977.

[7] Herold R, Wahl M, Regner G, et al. Thermodynamic benefits of opposed-piston two-stroke engines, SAE paper 2011-01-216[R]. Detroit: Society of Automotive Engineers, 2011.

[8] Zhou Longbao. Principle in internal combustion engine[M]. Beijing： China Machine Press，2003:48-79. (in Chinese)

[9] Matthew G, McGough E, Robert F. Experimental investigation of the scavenging performance of a two-stroke opposed-piston diesel tank engine, SAE paper 2004-01-1591[R]. Detroit: Society of Automotive Engineers,2004.

(Edited by Cai Jianying)

10.15918/j.jbit1004-0579.201524.0109

TK 421.8 Document code: A Article ID: 1004- 0579(2015)01- 0058- 07

Received 2013-12- 13

Supported by the National Natural Science Foundation of China (B2220110005)

E-mail: clzhao@bit.edu.cn