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Lowering Friction in timing chain drive systems

DSM Engineering Plastics outlines the importance of timing chain drive systems in the overall role of the IC engine, and as regulators continue to press for more efficient engines timing chains are the one area that can yield further gains

DSM Engineering Plastics


An important route to making cars more energy efficient is the development of more efficient engines. One way to improve engine efficiency is the reduction of friction. Sliding friction within the valve train drive system can be reduced by tuning the intrinsic tribological properties of plastics used for chain tensioner arms and guides. Within valve train drive systems, timing chains are used in roughly half the engines that come to the market in Europe.

To maintain the chain tension throughout the engine’s life, a tensioner arm is used at the slack side of the chain. In addition, guides and tensioner arms are used to prevent lateral resonance in the chain. A timing chain system typically incorporates one tensioner arm and three guides. A chain tensioner arm is used at the slack side of the chain to maintain the tension throughout the engine’s life. Chain guides are installed between sprockets, to prevent lateral resonance in the chain parts.

The tribology of these contacts is not only determined by the chain, tensioner arm, and guide surfaces, but also by the engine oil. This is sprayed onto the chain to reduce friction due to chain articulation, and migrates to the tensioner arm and guide contacts as well.

The friction depends on conditions such as chain tension, temperature, and sliding velocity. Friction under lubricated conditions is commonly represented by a “Stribeck” curve, which is schematically shown in Figure 1. (Below)

Figure 1. Schematic Stribeck curve for a sliding contact between two surfaces lubricated by a liquid. The dynamic coefficient of friction (μ), which is defined as the friction force at the contact divided by the normal force, is plotted on the vertical axis. On the horizontal axis is the Hersey number; this is the product of sliding velocity (v) and oil viscosity (η) divided by normal load (p).

When the sliding velocity is low, the pressure in the oil film is unable to support the normal load induced by the chain force and asperities on both surfaces are in direct contact with each other. (Asperities are the sharp, rough projections seen on an atomic scale on even those surfaces polished to a mirror finish.) Friction characteristics in this regime are quite similar to those of friction under dry conditions, although friction levels are generally much lower. This domain is termed the boundary lubrication regime.

When the sliding velocity increases, the pressure in the film increases and the film is able to support a larger portion of the normal load. The distance between the surfaces increases and asperities are gradually separated from each other. As a consequence, friction drops. The overall measured friction is now a combination of asperity-to-asperity friction and lubricant film friction. The velocity at which friction starts to drop depends on numerous parameters. This velocity domain is termed the mixed lubrication regime.

Further increasing the sliding velocity leads to a further increase of oil film pressure, enabling it to support the full normal load of the chain and separating the surfaces completely. At even higher velocities, the friction increases again due to the viscous drag in the oil. This velocity domain is termed the (elasto-)hydrodynamic regime. In this regime, the main friction determining parameters are the oil viscosity and the moduli of the tensioner arm, the guide, and the chain.

It should be noted that friction in a real timing chain drive system is far more complex than that which can be observed in lab tests. The conditions along the chain tensioner arm and guide may vary, for example. At sufficiently high sliding velocity, friction may be in the mixed lubrication regime at one location and in the boundary lubrication regime in another. The total friction measured over the full chain drive system is a weighed sum of the many individual contributions at all locations. This leads to a very complex relation to parameters such as chain tension, sliding velocity and temperature.

Figure 2. Schematic representation of the coefficient of friction in a chain to tensioner arm contact.





Underlying physics determining friction between engine oil lubricated steel and plastic

Various studies have been carried out for identifying the sensitivity of a number of parameters on the measured friction. One test used is known as the ball-on-pyramid test.

1. Oil Viscosity
With increasing viscosity, it becomes easier to separate the two surfaces by a lubricating film due to the higher film pressure. The transitions from boundary lubrication to mixed lubrication and to hydrodynamic lubrication shifts to lower sliding velocities with increasing oil viscosity. In the hydrodynamic regime, the coefficient of friction increases more steeply for higher viscosity liquids (higher viscous shearing).

2. Surface Roughness
For obtaining low values of the coefficient of friction, it is important to avoid direct contact between the two surfaces. The influence of the roughness of the plastic surface is shown in Figure 3. Increasing the surface roughness shifts the Stribeck curve in the direction of higher sliding velocity. In the boundary lubrication regime at the lowest sliding velocities, the coefficient of friction is lower for the sample with higher surface roughness.

Figure 3. Stribeck curves for steel ball in contact with neat PA46 samples of different surface roughness under otherwise identical conditions (1N normal load, 90°C, Castrol Edge 5W30).

In order to avoid direct contact between the two surfaces, the film thickness required to carry the normal load must be larger than the typical height of the asperity peaks on the surfaces. At around 1m/s sliding velocity, the required film thickness is on the order of a few tenths of a micron, which is on the same order of magnitude as the typical combined surface roughness for the smooth sample and the steel ball. For the rough sample, a minimum film thickness of at least a few microns would be needed, which is reached at sliding velocities of over 10 m/s.

3. Oil Additives
Modern engine oils are complex fluids with well-engineered additive packages that interact with surfaces to lower wear and friction, modify viscosity, and inhibit oxidation. The influence of the oil on the measured coefficient of friction is shown in Figure 4.

Figure 4. Coefficient of friction measured for a fresh engine oil (Castrol Edge 5W30) a used engine oil (Castrol Edge 5W30, 20000 km in a gasoline engine) and a mineral oil [9] all at 90°C. The plastic surface is HGR2 (PA46 with small amounts of friction modifiers).

When comparing the fresh engine oil to the mineral oil, there is a large difference in coefficient of friction, in particular in the boundary and mixed lubrication regime where there is contact between asperities on the two surfaces. The difference is attributed to the additive package in the engine oil interacting with the metal and/or the plastic surface.
The coefficient of friction for the used engine oil is significantly higher than that of the fresh engine oil. This may be because soot particles adsorb the additives, negatively affecting the effectiveness of the lubricant.

4. Temperature
For the base mineral oil without additives, the coefficient of friction in the boundary lubrication regime steeply increases with temperature. One possible explanation is that the lower modulus of the plastic at higher temperatures increases the true contact area and thereby the coefficient of friction. It could also be due to the lower viscosity of the oil.

Lowering Friction under Engine Oil Lubricated Conditions
Various modified PA46 materials engineered for lowering friction in particular in the boundary lubrication regime have been screened, with a focus on performance assessment of several friction-lowering additives and their possible synergistic effects. This resulted in Stanyl HGR2, a PA46 base material with a combination of friction modifying additives, including PTFE. The ball-on-pyramid test results for this new material and the standard PA66 and PA46 is shown in Figure 5. A significant reduction in friction has been achieved in the boundary lubrication regime. Results of tests on a motorized engine are shown in Figure 6.

Figure 5. Comparison of the performance of PA66 (Akulon S240-CH), PA46 (Stanyl TW341), and Stanyl HGR2 (PA46 with small amounts of friction modifiers) in Castrol Edge 5W30 engine oil at 90°C.


Figure 6. Results of motorized engine tests performed by BorgWarner Inc. on a Ford gasoline engine. All tests were performed using a 5W-20 engine oil heated to 93°C (± 2°C). The graph shows the influence of material on crankshaft torque: PA66, PA46 (Stanyl TW341), and Stanyl HGR2, with HGR2 showing the lowest torque of all.


17 May 2017


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