The liner (10) was fixed to the base (4). A piston ring holder was attached to the machine head (8) that applied the movement and load. The ring (11) was attached to the holder (8). This configuration ensures good line contact between the bodies under applied loads. A more detailed description of this setup was given elsewhere    [13].

The stroke of reciprocating motion was set 4 mm that, in combina- tion with the motion frequency of 25 Hz, is representative for simula- tion of TDC friction conditions (see Fig. 2)  for  90 mm  stroke  and 1000 rpm. This yielded speed of the piston 0.1 m/s at 2 mm from the top dead centre.

The normal load was set 616 N to get the calculated contact pressure 20 MPa, same as in the real ring – liner contact. This load was much below the scuffing load as determined in separate extreme pressure tests conducted for the chosen lubricants following ASTM D5706 standard. As we pursued simulation of mixed/boundary lubri- cation regime, oil volume was limited to 1 ml. From our previous tests and the present results, this volume of oil was found sufficient for the whole test duration of 12 h. The temperature during the test main- tained at 180 °C using a resistive heater. At least two repetitions of each test were done. After the tests, the wear scars were examined by optical and confocal microscopy. Wear was evaluated by mass method using a microbalance with resolution of 10 μg. Mass method was chosen instead of the volume one because of curved macrogeometry and irregular microgeometry of contact surfaces which make calculation of worn volume difficult and increase uncertainty. The combined uncer- tainties of the reported values were 0.01 for friction and 0.1 mg for mass wear.

Chemical composition of some of the worn surfaces was analysed using X-ray Photoelectron Spectrometry (XPS). The angle between the hemispherical analyser (Specs-PHOIBOS100) and the surface was 60°. An X-ray radiation source with MgKa line (1253.6 eV) was used. The survey spectra were obtained with electron energy step of 0.25 eV and pass energy 40 eV. Fine core level spectra were measured with electron energy step 0.1 eV and pass energy 15 eV. Before the analysis of the obtained data the contribution of the MgKa satellite line was subtracted and  the  spectra  were  subjected  to  Shirley  background    subtraction Fig. 2. Friction coefficient evolution during the engine cycle: a) the total stroke of 96 mm stroke; b) the zones of interest for tribological simulation; c) friction coefficient evolution during simulated tribo-test with 4 mm   stroke.formalism [29].

2.2. Materials

The base materials were different types of cast iron for the ring  and

Table 1

Microhardness of piston rings with various surface    coatings.

Coating Hardness (GPa) Roughness  Ra (µm)

DLC 30.4 0.165

MoS2 6.8 0.167

study the effect of surface micro-geometry on tribological performance at TDC. The texture and roughness of honing is believed to be important for  oil  retaining  and  distribution  on  the  surface  [30]. Fig. 3 shows reconstructed 3D images of the samples. Grey scale defines heights of surface  features.

In addition, several coated piston rings were tested against un- coated cylinder liner. The coatings were produced by Physical Vapour Deposition technique at IK4-TEKNIKER or by a third company. The most relevant parameters for tribological performance – microhard- ness and roughness – are shown in Table 1. Surface roughness of the coatings was measured using confocal microscopy as well as contact profilometer. The main difference between the two CrN coatings is the increased thickness for CrN_2 up to 20 µm. Surface roughness was not significantly affected by coating deposition and remained very similar to that of the bare  rings. 活塞环/气缸套在摩擦和磨损英文文献和中文翻译(3):http://www.youerw.com/fanyi/lunwen_204657.html