aEcole Supérieure en Génie Electrique et Energétique - ESGEE, Oran, Algeria
a,bLaboratory of Gas Combustion and Environment, Department of Mechanical Engineering, University of Sciences and the Technology of Oran, L.P 1505 El -Menaouer, USTO 31000 Oran, Algeria
( *Corresponding author: acer.msn@hotmail.fr)
ABSTRACTFor different operating conditions of an internal combustion engine, the piston-ring-liner compartment represents one of the largest sources of friction and power losses. The present paper is a contribution to evaluate the effect of the compression ring profile on the main tribological performance of the lubricant in a four-stroke diesel engine. A one-dimensional analysis was developed using GT-Suite software for the hydrodynamic lubrication between the compression piston ring and the cylinder wall. A numerical method was applied to analyze the influence of different ring geometrical designs during the working cycle on oil film thickness, frictional force, power losses. The Takigushi simulation model was used to validate and verify the developed numerical model. The results obtained showed a relatively good qualitative and quantitative compatibility between the two applied models. The developed numerical model allows the investigation of the influence of ring design parameters to improve the tribological performance of piston engines. |
RESUMENEstudio numérico del efecto de perfil del primer anillo del pistón de compresión sobre el rendimiento hidrodinámico del lubricante de un motor diésel. El comportamiento del revestimiento del anillo del pistón de combustión representa una de las mayores fuentes de fricción y de potencia para diferentes condiciones de funcionamiento de un motor de combustión interna. El presente trabajo evalúa el efecto del perfil del anillo de compresión sobre el principal comportamiento tribológico del lubricante en un motor diésel de cuatro tiempos. Se desarrolló un análisis unidimensional utilizando el software GT-Suite para la lubricación hidrodinámica entre el segmento del pistón de compresión y la pared del cilindro. Se aplicó un método numérico para analizar la influencia de los diferentes diseños geométricos de anillos durante el ciclo de trabajo sobre el espesor de la película de aceite, la fuerza de fricción y la pérdida de potencia. Se utilizó el modelo de simulación de Takigushi para validar y verificar el modelo numérico utilizado. Los resultados obtenidos mostraron una compatibilidad cualitativa y cuantitativa buena entre los dos modelos aplicados. El modelo numérico utilizado permite analizar la influencia de los parámetros de diseño de anillos para mejorar el rendimiento tribológico de los motores de pistón. |
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Submitted: 5 March 2020; Accepted: 16 July 2020; Available On-line: 20 October 2020 Citation/Citar como: Menacer, B.; Bouchetara, M. (2020). “Numerical study of the effect of first ring profile type on hydrodynamic performance of the lubricant in diesel engine”. Rev. Metal. 56(3): e177. https://doi.org/10.3989/revmetalm.177 KEYWORDS: Cylinder liner; Diesel engine; Gas dynamics; Hydrodynamic lubrication; Oil film thickness; Piston ring; Power losses PALABRAS CLAVE: Anillo de pistón; Camisa de cilindro; Dinámica de gases; Espesor de la película de aceite; Lubricación hidrodinámica; Motor diésel; Pérdida de energía ORCID ID: Brahim Menacer (https://orcid.org/0000-0002-7922-1847); Mostefa Bouchetara (https://orcid.org/0000-0001-6546-9086) Copyright: © 2020 CSIC. This is an open-access article distributed under the terms of the Creative Commons Attribution 4.0 International (CC BY 4.0) License. |
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The optimisation of the piston rings as important components of combustion engines can contribute to the engine friction reduction, since about 20% to 40% of the friction losses of the engine are due to the piston ring assembly. The main functions of the piston rings are the sealing between the combustion chamber and the engine crankcase, the support for the heat transfer from the piston to the cylinder wall and the control of the engine oil consumption. During the compression and working strokes, the compression ring seals the combustion gases and prevents blow-by.
Many researchers have tried to understand the mechanism of piston ring motion through numerical and experimental investigations. Radakovic and Khonsari (1997) presented the governing equations and the appropriate numerical solution method to treat thermohydrodynamic problems implying thin-film flows in the presence of transversal compression and shear thinning. They also analyzed thermal and shear effects of oil film on ring performance. Takiguchi et al. (2003) studied the variation of the oil film thickness of the top and second piston ring of a truck diesel engine during a cycle by means of clearance capacitive sensors in the sliding surfaces of the rings. For the investigation of tribological characteristics, (Chaudhari and Sutaria, 2016) compared the results obtained by numerical modelling of the frictional force of the various friction sources with the experimental results. Srinadh and Rajasekhara (2015) designed three piston rings with different profiles for a diesel engine using Pro/Engineer software and analyzed the stresses and displacements of the piston and piston rings. The model used was the same for each mechanism, but with different procedures and assumptions. However, they did not specify the steps in the simulation.
The objective of this work is to study the effect of three piston ring profiles on tribological parameters such as oil film thickness, ring friction and power loss by using of the GT-Suite simulation software. The oil film thickness was calculated from the theoretical analysis and compared with the results of Takiguchi et al. (2003). The main aim is to identify the best ring variant that would reduce friction and contribute to improved engine performance.
The in-cylinder pressure which is one of the inputs for the simulated engine code was modulated using the first law of thermodynamics (Sonthalia and Kumar, 2013). For the simulation model, we take into account the following assumptions (Notay et al., 2019):
| – | The friction between the rings and the liner includes the lubricant shear friction given by the Reynolds equation and the friction due to the Greenwood-Tripp asperity contact. |
| – | In the piston ring modeling, we have the moments and forces for each ring. The hydrodynamic pressure distribution between running surface of the ring and the liner is determined by solving the Reynolds equation in each time step (Usman and Park, 2016). |
| – | The volumes are connected due to the actual clearances of ring end gaps and the actual position of the rings in the grooves. The gas flow behind the rings and between ring and groove flanks is considered. |
| – | The oil film is taken into account between the ring running surface and liner by calculating the pressure distribution in the clearance. |
| – | In this study, lateral pressure was neglected because only the ring-piston-liner interaction is of interest. The installed ring is in a compressed state to adapt to the cylinder (Wakuri et al., 1995). |
| – | It was assumed that the ring contracts and extends uniformly around the circumference and that the ring sits on a single point with a running contact on the top or bottom of the groove. The contact points also seal the pressure on one side from the other, so that there is a gradual change in pressure at this contact point. |
| – | The lubrication regime in the cylinder is influenced by the variation in piston speed, which affects the friction between the piston ring and the liner throughout the piston stroke. In this study, we consider hydrodynamic lubrication, which means that the ring always moves over a complete fluid film. |
| – | The frictional force reaches its maximum where the piston speed is highest, i.e. the piston position is at mid-stroke. If the engine speed or oil viscosity is high, a thick oil film is formed which is not completely ejected even at dead centers where the piston speed is zero (Wannatong et al., 2008). |
| – | The effect of temperature on the oil viscosity has also not been taken into account. |
Due to the high inertia of the connecting rod - crank system, the position of the piston will not be affected by the piston ring action. The axial position x, speed U and acceleration a of the piston relative to the liner as a function of crankshaft angular position, measured from TDC, are given respectively by (Tian, 1997):
R is the crank radius (m), L is distance between bearing centers on the connecting-rod and the crankshaft lies in the same plane as the piston motion (m), ω crankshaft angular velocity (1/s) and t is time (s).
In the case of hydrodynamic lubrication, the oil fully supports the load of the ring on the liner, and therefore the frictional force generated by the segment-liner interaction depends on the lubricant properties, the height and width of the film below the ring surface. To determine the frictional force, coupled equations governing the fluid mechanics of the lubricant and the forces acting on the segment have to be solved. Figure 1 shows the schematic of studied system. In the general case, the unknowns are the minimum oil film thickness c, the input wetting coordinate x1 and the output wetting coordinate x2. Between both coordinates, there is a transition area where the oil attaches and detaches from the ring. Although this zone has a very small axial width compared to the axial width of the ring, it has an important function in the lubrication of the system. W, the net resultant radial force (N), a the width of parabolic portion of the ring face (mm), b the width of the straight portion of the ring face (mm); δ the inclination of the wedge (°) and c minimum film thickness (mm).
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Reynolds’ equation establishes the relationship between the height, width and the oil film thickness between ring and liner and the oil pressure gradient. In the present study, a 2-D approach is used in which the ring and liner lubrication parameters are determined as a function of piston position, and the Reynolds equation is thus reduced to a 1-D form. Under these conditions, the Reynolds equation is written (Cho and Jang, 2004; Delprete and Razavykia, 2018; Delprete et al., 2019):
h is the oil film thickness under the ring as function of the coordinate distance calculated from the leading edge of the ring x (mm), U the piston velocity and η lubricant viscosity (cst).
Taking into account the symmetry of the axes along the cylinder axis, equation (4) becomes (Balakrishnan and Rahnejat, 2005):
The variation of the film thickness under the ring with respect to position and time is (Morris et al., 2013):
where, hp is the ring profile height (m) and hr is the ring reference distance (m).
The surface shear stress of lubricant between two parallel plates is given by the following relationship:
The friction force can be obtained by integration of Eq. (7):
The power loss is given by the following equation (Jeng, 1992):
The Reynolds’s equation (5) is solved by using the finite difference method and the application of the above approximations with boundary conditions. Relaxation method is used for finalizing the numerical solution using an initial guess for film thickness equation (6). The Newtonian lubricant is used in the numerical scheme for the lubrication solution. The average Reynolds equation is solved to calculate the distribution of hydrodynamic pressures. The gradients of pressure depend on the film thicknesses. Gauss Seidel iterative numerical scheme is used to solve the Reynolds and the oil film thickness equations simultaneously and show maximum and minimum film thicknesses and pressure fields.
The contact between the ring and cylinder wall is ensured by the spring action inherent in the ring itself, which expands the ring radially (Zavos and Nikolakopoulos, 2015). The ring can be inclined or moved axially up and down in the groove. Here, only the axial movement of the segment in the groove is taken into account and it is assumed that the surfaces of the groove are flat. The forces acting on the ring in the axial direction are shown in Fig. 2 and are expressed as follows (Mohamad et al., 2015).
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The pressure force acting on the ring is given by (Gulwadi, 2000):
Ar is the ring area in the radial direction, Pi the pressure force in the regions i = 1; 2; 3.
The friction force is calculated from the relation (Guo et al., 2011):
with:
P is the pressure behind the ring, dr the diameter of the ring , Tr the thickness of the ring, U the piston speed in m/s and μ the oil viscosity in cst.
The inertia force is due to the mass of the ring and calculated from the relation:
Mr is the ring mass (kg),a the piston acceleration (m/s2).
To solve the Reynolds equation, the boundary conditions,P(x = –b) = P1(t), P(x = a)=P2(t) and the piston position during the engine operating cycle must be taken into account, Fig. 3. The vertical velocity of the body VB = dc/dt is determined by the balance between the liner of the ring force W and the lifting force due to the lubricant pressure, expressed by (Soualmia and Bouchetara, 2015):
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The compression ring profile play an important role in the entertainment of the lubricant between the top piston and cylinder liner. Figure 4 and table 1 show the different top ring profiles and its properties used in this study. The cylindrical shape of the ring provides better hydrodynamic lubrication conditions and the axially shorter contact surface to the liner surface improves sealing. In addition, the negative effects of cylinder deformations during engine operation can be better compensated. Piston rings of this profile type are actually used as compression rings.
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| Property | Type I | Type II | Type III |
| Ring Surface Upper Profile Height | 1.5 μm | 3.5 μm | 6.5 μm |
| Ring Surface Lower Profile Height | 1.5 μm | 3.5 μm | 6.5 μm |
| Ring Surface Crest Position | 5 mm | 5 mm | 5 mm |
Table 2 presents the data used in this investigation. The determination of tribological performance was carried out under full load of a diesel engine and at an engine speed of 2000 rpm.
| Parameter | Value | Units |
| Connecting Rod Length | 127 | mm |
| Crank radius | 31.5 | mm |
| Ring Axial Width | 1.63 | mm |
| Ring Radial Width | 2.63 | mm |
| Bore radius | 48 | mm |
| Ring Density | 7900 | kg·m−3 |
| Ring Elastic Pressure | 93539 | N·m−2 |
| Ring Modulus of Elasticity | 2.05 e11 | Pa |
| Engine speed | 2000 | rpm |
| Oil Density | 881.5 | kg·m−3 |
| Oil Viscosity | 0.008736 | Pa-s |
The speed piston can be determined by deriving the equation that calculates the distance between the piston and top dead center. Figure 5 shows the variation of piston position and speed as a function of crankshaft angle at 2000 rpm for a diesel engine. Piston ring lift is due to piston acceleration, which is the derivative of the piston speed equation, Fig. 6.
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Figure 7 shows the variation of hydrodynamic power losses versus crank angle for different compression ring profile at full load and at 2000 rpm engine speed. During the operating cycle of the engine, the maximum power losses for all three ring variants are all reached at mid-stroke of the piston and then cancel each other out at TDC and TDC, corresponding to the piston velocity variation, Fig. 5. The oil film thickness and the hydrodynamic shear increase also at mid-stroke.
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Figure 8 present the variation in asperity friction power loss during a full load engine cycle at 2000 rpm. The maximum values are in the expansion phase and minimum values in the other phases of the engine cycle. It can be attributed to the fact that there is an increase in asperity contact near the TDC and BDC due to the mixed lubrication regime, whereas hydrodynamic lubrication exists for most part of the stroke. The type 03 of ring design presents the higher values.
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Figure 9 shows the variation of hydrodynamic friction force versus crank angle for different ring profiles. That maximum friction value is at point of maximum cylinder pressure. The negative part of the curve is due to the change of piston speed direction through the reciprocating motion. It has been observed that the highest friction force value occurred during the expansion stroke (from -110° to 180°) with a peak of 235 N just after TDC firing. The friction was found maximum for type 1 ring profile, on the other hand type 3 has minimum friction force. At TDC the film thickness for type 3 profile is found to be lowest. It is expected that the friction due to lubricant viscosity was minimum.
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In general, the maximum asperity frictional force is reached at TDC and BDC during the operating cycle, due to the low critical speed and, consequently, the low oil flow at these points. Figure 10 shows the asperity frictional force between the upper ring and the cylinder liner as a function of the crankshaft angle for different ring profiles at an engine speed of 2,500 rpm and full load. Type 1 has the lowest asperity friction force, followed by type 2 and the highest force with type 3 especially at TDC after the compression stroke. Figure 11 shows the variation of the maximum contact oil pressure as a function of the crank angle for different ring profiles. The maximum value of the hydrodynamic oil pressure is reached near the TDC. Ring profile type 3 has the maximum oil pressure, followed by type 2 and type 3 with the minimum oil pressure. Figure 12 shows the variation in film thickness for different types of ring profile. The Takiguchi model (Takiguchi et al., 2003) was used to validate the results obtained by the developed model. The film thickness is maximum at the mid-point of the different cycle phases. In the expansion stroke, the oil film thickness is lower. At the dead centers, the film is very thin, especially at the BDC after the expansion stroke and at the TDC at the beginning of the intake stroke, which can lead to high wear of the cylinder wall due to contact between the surfaces. It can be observed that type 01 of the ring face profile provides a thicker film than the other two variants.
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Figure 13 shows the variation in shear rate as a function of crank angle for different compression ring profiles. The maximum shear rates vary over the cycle from 1.5.10−6 to 3.0.10−6 s−1. Type 1 has a higher effective oil film shear rate, followed by Type 02 and the lowest with Type 3.
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Figure 14 shows the variation of friction tension as a function of crank angle for different compression ring profiles. In the combustion and expansion period, there is a noticeable change in the frictional tension. The variant of type1 has higher frictional tensions than the other variants. Figure 15 illustrates the gap variation versus the crank angle for different ring profiles. A remarkable change in the gap is observed in the combustion and expansion period.
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The theoretical results obtained with the numerical model developed prove that the profile of the compression rings has an important effect on the tribological behavior of the system piston ring-cylinder liner. The modification of the compression ring profile contributes to the reduction of frictional power losses and thus to the improvement of the tribological performance of the internal combustion engine. The increase in asperity contact causing low oil films at TDC and BDC as well as friction asperity power losses can be observed. At mid-stroke of the piston motion, thicker oil films were observed despite the presence of higher hydrodynamic frictional forces and lower oil viscosity. There was a good agreement between the results obtained with the developed numerical model and those of the Tagukishi model. In the present study, the following results could be deduced:
| – | The lowest frictional force is for the type 3 ring profile, which results in a reduction in engine friction losses and an increase in lubricant and fuel economy. |
| – | The minimum oil film thickness for type 3 compression ring profile does not necessarily mean a higher frictional force and shear stress. Despite a minimum oil film thickness, the wear rate is not affected. |
| – | In future work, it would be advisable to take the blow-by phenomenon into account and analyze its effect on the ring motion in the piston groove and on the tribological behavior of the piston-ring-cylinder liner assembly. |
| ○ | Balakrishnan, S., Rahnejat, H. (2005). Isothermal transient analysis of piston skirt-to-cylinder wall contacts under combined axial, lateral and tilting motion. J. Phys. D Appl. Phys. 38 (5), 787–799. https://doi.org/10.1088/0022-3727/38/5/018. |
| ○ | Chaudhari, T., Sutaria, B. (2016). Investigation of friction characteristics in segmented piston ring liner assembly of IC engine. Perspect. Sci. 8, 599–602. https://doi.org/10.1016/j.pisc.2016.06.032. |
| ○ | Cho, J., Jang, S. (2004). Effects of skirt profiles on the piston secondary movements by the lubrication behaviors. Int. J. Auto. Techn. 5 (1), 23–31. |
| ○ | Delprete, C., Razavykia, A. (2018). Piston ring–liner lubrication and tribological performance evaluation: A review. J. Eng. Tribol. 232 (2), 193–209. https://doi.org/10.1177/1350650117706269. |
| ○ | Delprete, C., Razavykia, A., Baldissera, P. (2019). Detailed analysis of piston secondary motion and tribological performance. Int. J. Engine Res. 21 (9), 1647–1661. https://doi.org/10.1177/1468087419833883. |
| ○ | Gulwadi, S.D. (2000). Analysis of tribological performance of a piston ring pack. Tribol. T. 43 (2), 151–162. https://doi.org/10.1080/10402000008982325. |
| ○ | Guo, Y.B., Lu, X.Q., Zou, D.Q., He, T., Li, W.Y. (2011). Analysis of Piston Ring Lubrication with Different Lubricant Supply. Adv. Mat. Res. 200, 700–706. https://doi.org/10.4028/www.scientific.net/AMR.199-200.700. |
| ○ | Jeng, Y.R. (1992). Theoretical analysis of piston-ring lubrication. Part I—fully flooded lubrication. Tribol. T. 35 (4), 696–706. https://doi.org/10.1080/10402009208982174. |
| ○ | Mohamad, S.A., Lu, X., Zheng, Q. (2015). Numerical modeling of lubrication of piston ring of two-stroke marine diesel engine considering the effect of multi-scale grooves on the cylinder liner. P. I. Mech. Eng. J.-J. Eng. 229 (8), 989–1002. https://doi.org/10.1177/1350650114556400. |
| ○ | Morris, N., Rahmani, R., Rahnejat, H., King, P.D., Fitzsimons, B. (2013). The influence of piston ring geometry and topography on friction. P. I. Mech. Eng. J.-J. Eng. 227 (2), 141–153. https://doi.org/10.1177/1350650112463534. |
| ○ | Notay, R.S., Priest, M., Fox, M.F. (2019). The influence of lubricant degradation on measured piston ring film thickness in a fired gasoline reciprocating engine. Tribol. Int. 129, 112–123. https://doi.org/10.1016/j.triboint.2018.07.002. |
| ○ | Radakovic, D.J., Khonsari, M.M. (1997). Heat Transfer in a Tin-Film Flow in the Presence of Squeeze and Shear Thinning; Application to Piston Rings. J. Heat Transfer. 119 (2), 249–257. https://doi.org/10.1115/1.2824217. |
| ○ | Sonthalia, A., Kumar, C.R. (2013). The effect of compression ring profile on the friction force in an internal combustion engine. Tribology in Industry 35 (1), 74–83. |
| ○ | Soualmia, A., Bouchetara, M. (2015). Numerical modeling of the hydrodynamic behavior of the couple piston liner. Mechanics 21 (6), 450–456. https://doi.org/10.5755/j01.mech.21.6.7035. |
| ○ | Srinadh, M., Rajasekhara, K.B., (2015). Static and Thermal Analysis of Piston and Piston Rings. IJEMAS 3 (8), 51–58. |
| ○ | Takiguchi, M., Harigaya, Y., Suzuki, M. (2003). Analysis of oil film thickness on a piston ring of diesel engine: Effect of oil film temperature. J. Eng. Gas Turbines Power 125 (2), 596–603. https://doi.org/10.1115/1.1501078. |
| ○ | Tian, T. (1997). Modelling the performance of the piston ring pack in internal combustion engines. Thesis Ph. D., Massachusetts Institute of Technology. |
| ○ | Usman, A., Park, C.W. (2016). Optimizing the tribological performance of textured piston ring–liner contact for reduced frictional losses in SI engine: Warm operating conditions. Tribol. Int. 99, 224–236. https://doi.org/10.1016/j.triboint.2016.03.030. |
| ○ | Wakuri, Y., Soejima, M., Ejima, Y., Hamatake, T., Kitahara, T. (1995). Studies on friction characteristics of reciprocating engines. SAE International (Technical Paper 952471). https://doi.org/10.4271/952471. |
| ○ | Wannatong, K., Chanchaona, S., Sanitjai, S. (2008). Simulation algorithm for piston ring dynamics. Simul. Model. Pract. Th. 16 (1), 127–146. https://doi.org/10.1016/j.simpat.2007.11.004. |
| ○ | Zavos, A.B., Nikolakopoulos, P.G. (2015). Simulation of piston ring tribology with surface texturing for internal combustion engines. Lubr. Sci. 27 (3), 151–176. https://doi.org/10.1002/ls.1261. |