INFLUENCE OF DIFFERENT SHAFT SURFACE FINISHES ...irf/Proceedings_IRF2020/data/...3760 (1996) and DIN...

Proceedings IRF2020: 7th International Conference Integrity-Reliability-Failure.

J.F. Silva Gomes and S.A. Meguid (editors), INEGI-FEUP (2020), pp.351-368.

https://paginas.fe.up.pt/~irf/Proceedings_IRF2020/

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PAPER REF: 17206

INFLUENCE OF DIFFERENT SHAFT SURFACE FINISHES ON THE LUBRICATION AND FRICTION CONDITION OF ELASTOMERIC LIP SEALS Markus Schulz(*), Matthias Baumann, Frank Bauer

Institute of Machine Components (IMA), University of Stuttgart, Stuttgart, Germany (*)Email: markus.schulz@ima.uni-stuttgart.de ABSTRACT

The duty parameter diagram is a common tool to compare and analyse the lubrication and friction condition of different radial shaft sealing systems. In addition, it also allows an evaluation of the suitability of differently manufactured sealing counterfaces. This paper deals with the influences of different shaft surface structures on the formation of a lubricating film and a sealing gap. For this purpose, friction torque measurements were carried out, starting from very low sliding speeds of about 0.0042 m/s (corresponds to 1 rpm at a shaft diameter of 80 mm) up to 10.5 m/s. The aim is to cover the mixed friction area in addition to the liquid friction area and thus determine the position of the transition point between both friction states.

Keywords: elastomeric lip seals, lubrication, friction, duty parameter, belt grinding, superfinishing.

INTRODUCTION

Elastomeric lip seals are often used to seal a shaft entry or outlet point. The sealing system is a tribological system and consists therefore not only of the elastomeric lip seal itself, but also includes the shaft and the fluid to be sealed. The surface structure of the shaft counterface influences the lubrication and friction condition of the sealing system. In order to reduce the costs of shaft manufacturing, the standard manufacturing process plunge grinding, is more and more often replaced by alternative, faster and cheaper manufacturing processes like belt grinding or superfinishing. Due to the replacement of the manufacturing process, different surface structures arise. These structures directly influence the sealing mechanism as well as the lubrication and friction condition of the sealing system. As a result, often problems with increased wear or leakage occur.

Reason for the above is a lack of empirical experience with use of such manufactured shafts in radial shaft seals. The roughness values of a shaft surface required in various standards (DIN 3760 (1996) and DIN 3761 (1984)), for example, are actually only valid for plunge ground shafts, because they are empirical data, which were collected over years with sealing systems with plunge ground shafts. Aim of this paper is therefore to collect empirical values regarding friction caused by the use of belt ground and superfinished shafts in radial shaft sealing systems and to draw conclusions about the roughness of the shaft surface to be maintained. For this reason, sealing systems each with three belt ground (BG) and superfinished (S) shafts are compared with a reference system with a plunge ground shaft.

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THE TRIBOLOGICAL SYSTEM OF RADIAL SHAFT SEALS

A radial shaft sealing system is a tribological system and consists, in addition to the sealing element, of the shaft counterface, the fluid to be sealed and the ambient conditions, as shown in Figure 1.

Fig. 1 - Tribological system of a radial shaft seal according to (Bauer, 2019)

Studies of Hermann and Seffler (1985) as well as of Kawahara and Abe (1978) have shown that the functional principle of radial shaft seals is based on an active sealing mechanism that actively conveys fluid from the air- to the oilside of the seal. This sealing mechanism prevents leakage and provides a fluid exchange, which leads to a heat dissipation in the contact area between the sealing edge and the shaft counterface. As an explanation for this sealing mechanism, various function hypotheses/ principles have been established. The three main principles (distortion (Kammüller, 1986) and wiping edge principle (Jenisch, 1991) as well as the side flow hypothesis (Kammüller et al., 1986a, 1986b), overview in (Müller et al., 1998)) base on an asymmetrical pressure distribution in the sealing contact due to asymmetric flank angles on the air- and the oilside of the sealing edge. Based on the assumption that surface structures of a shaft counterface wear out in operation (Horve, 1991; Jagger et al., 1966), only the roughness of a sealing edge was considered in these principles. However, it was possible to prove that no or hardly any wear occurs on a shaft counterface when using a radial shaft seal made of nitrile butadiene rubber (NBR) (Schuler, 2014). For this reason, the roughness and surface structures of a shaft counterface must also be considered with regard to the sealing mechanism of a radial shaft seal. Due to the manufacturing, on shaft counterfaces occur isotropic or anisotropic surface structures. Most technical surfaces are anisotropic, i.e. they show in different directions different topographies. Often the structures on shaft counterfaces run in the circumferential direction of the shaft. If the structures have an angular deviation from the circumferential direction, they are called lead structures or just lead. Depending on the orientation of the lead and the direction of shaft rotation, the lead structures convey fluid in the direction of the air, or the oil side of the sealing, as shown in Figure 2. Thus, leakage or an insufficient lubricated sealing edge occur.

The fluid is not only axially deflected by the surface structures at the sealing edge and the shaft counterface, but a hydrodynamic pressure is also built up in the fluid due to the roughness structures of both surfaces and the resulting gap height change. If the hydrodynamic pressure reaches a higher value than the mean surface pressure pm between sealing edge and shaft counterface, the sealing edge lifts off the shaft counterface and a hydrodynamically lubricated sealing contact is created.

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Fig. 2 - Influence of lead on the function of a sealing system (Schulz et al., 2016)

Depending on the characteristics of the surface structures on the shaft counterface, the build-up of hydrodynamic pressure is supported to a greater or lesser extent, as can be seen in Figure 3. In the case of a plunge ground shaft counterface, the decrease in width and depth of the individual grinding grooves in circumferential direction supports the build-up of hydrodynamic pressure. In contrast to this, no hydrodynamic pressure is built up on a belt ground shaft counterface, with circumferential, closed grinding grooves, due to a lack of cross-sectional change at the grinding grooves.

Fig. 3 - Hydrodynamic Pressure build-up at a plunge ground groove (left) and a belt ground,

circumferential groove (right) according to (Bauer, 2019)

The lubrication condition of a radial shaft seal can be characterised by friction torque measurements and duty parameter diagrams. The duty parameter diagram is alike the Stribeck curve from field of hydrodynamic bearings, as shown in Figure 4.

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Fig. 4 - Duty parameter diagram for plunge ground shaft counterfaces (Jung, 2013) The duty parameter, G, is a dimensionless index to describe the hydrodynamic condition of a sealing system and is defined as

# = ∙ �[U (1)

with the dynamic viscosity η, the angular velocity ω and the mean surface pressure pm. In combination with the friction coefficient f and the duty parameter, the duty parameter diagram can be drawn out. As most variables of the duty parameter are temperature-dependent, particular attention was paid within this work to the dynamic viscosity η and the mean surface pressure pm (due to the temperature-dependence of the radial force). Both parameters were determined in dependence of the sealing gap temperature. Since the temperature in the sealing gap cannot be measured, it is possible to estimate this temperature using the friction torque caused by a seal with the program InsECT (Feldmeth et al., 2013). The coefficient of friction f is determined by the friction torque caused by a radial shaft seal and its radial force. The transition of mixed and liquid friction is in a range of G = 10-7 (Buhl, 2006; Jung, 2013). Depending on the manufacturing of the used shaft counterfaces the duty parameter diagram shows a different course (Jung, 2013). First results of our investigations with belt ground and superfinished shaft counterfaces were published in (Schulz et al., 2018a). The results show a dependence of the friction torque caused by the seal on the roughness of belt ground shaft counterfaces. With superfinished shaft counterfaces, a higher friction was determined than with belt or plunge ground shaft counterfaces. With exception of the finest superfinished shaft, the area of mixed friction and thus a transition point could not be proven in any test. For this reason, the only hypothesis that could be put forward was that with superfinished shaft counterfaces with decreasing roughness the transition point is shifted to larger duty parameters, and the course is restricted to a smaller range of duty parameters.

For these reasons, further tests were carried out under modified test conditions (speed collective, test duration, oil splashing instead of oil sump, ...) in order to be able to prove both types of friction in the duty parameter diagram and to confirm or falsify the hypothesis.

MANUFACTURING AND SURFACE ANALYSIS OF SHAFT COUNTERFACES

The shaft counterfaces used in this investigation were manufactured with a superfinish add-on device Supfina 210 shown in Figure 5, which allows beltgrinding and superfinishing. During manufacturing, a grinding belt is pressed via a pressure roller against the shaft counterface rotating on a lathe. The grinding belt is continuously moved forward by a driven role in order to keep new abrasive grains in engagement. Since the grinding belt is not axially fixed on the pressure roller, slight axial movements of the grinding belt may occur. In order to obtain

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differently rough shaft counterfaces, three grinding belts with different grain sizes (180, 500 and 1200) were used. Each shaft was exclusively manufactured with a grinding belt of one grit, in one clamping. Depending on whether or not the pressure roller carries out an axial movement, we speak of belt grinding (without axial movement) or superfinishing (with axial movement). In the following belt ground shafts are referred to as BG and superfinished shafts as S. The appended number refers to the grain size of the used grinding belt (1 ≙ 180, 2 ≙ 500 and 3 ≙ 1200). For comparison purposes, a plunge ground shaft counterface is also examined, this is referred to as reference. The surface analysis includes roughness and lead measurements as well as a visual analysis. The results of the roughness measurements are shown in Figure 6. The grey-shaded areas mark roughness limits specified in DIN 3760 (1996). These are only intended to provide a better assessment of the roughness values, as the limits were determined empirically using plunge ground shaft counterfaces and are therefore only valid for surfaces manufactured in this way. As can be seen from the graph, the values of the roughest belt ground shaft counterface is above the specified range for all three roughness parameters. In the case of superfinished shaft counterfaces, all roughness parameters Ra and Rz, with exception of the Rz value of shaft S1, are below the minimum required roughness values.

Fig. 6 - Roughness parameters of shaft counterfaces

All used shaft counterfaces were comprehensively examined for lead. However, since both macro and micro lead analysis are subject to certain limitations depending on the manufacturing process used for the shaft surfaces, the limitations and the respective results of the lead analyses are only briefly presented.

Macro lead analysis according to the CARMEN method shows a zero lead for all belt ground shaft counterfaces, which correlates in lead depth with the grain size of the used grinding belt. For the superfinished shaft counterfaces, the macro lead analysis shows a left-hand lead for shafts S2 and S3, while shaft S1 is macro lead-free. The validity of the macro-lead analysis for

Fig. 5- Superfinsish-add-on device Supfina 210

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superfinished shaft counterfaces is questionable, as it only indicates one (dominant) lead angle, whereas the surface contains two "equivalent" lead angles due to the crossed structures. A statement as to whether macro lead is present in such surfaces depends very much on the axial position of the sealing edge or the measuring range. Since a lead measurement cannot be carried out before every test at the exact position where the sealing edge will be in contact with the shaft counterface, it is not possible to evaluate the shafts on the basis of a macro-lead analysis. The reference shaft counterface used, which is plunge ground, is macro-lead free.

Micro lead was analysed by IMA-Microlead® Analysis developed at Institute of Machine Components (IMA) at University of Stuttgart. Figure 7 shows angle and volume distribution curves of the belt ground shaft counterface BG3 and the reference shaft. Angular and volume distributions of all belt ground shaft counterfaces are significantly narrower than those of the plunge ground reference shaft. Due to the very narrow distributions, a slight tilting or eccentricity of a shaft during its manufacturing or later in operation is sufficient either to create micro lead structures on shaft counterface or to tilt the structures running in circumferential direction, thus conveying fluid. All belt ground shaft counterfaces show their maximum at 0.01° in both distribution curves and thus exhibit a slight micro right-hand lead. The plunge ground reference shaft counterface has its maximum in the volume distribution curve at 0.05°. When applying the specifications of Institute of Machine Components (IMA) for an optimum shaft counterface (Universität Stuttgart, Institut für Maschinenelemente, 2020), such a surface can still be used as a shaft counterface. For superfinished shaft counterfaces with crossed structures, an evaluation of micro lead analysis is not possible, because the algorithm is not able to detect crossed structures so far.

Fig. 7 - Comparison of angle and volume distribution curves of belt ground shaft counterface BG3 and

plunge ground reference shaft counterface

Instead of the lead analysis of the shaft counterface the analysis of three-dimensional roughness parameters seems to be more suitable to describe and characterise alternatively manufactured

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shaft counterfaces. However, a correlation of three-dimensional roughness parameters with characteristic friction values has not yet been carried out, but is only planned in a subsequent step.

Figure 8 shows the surface structures that result from the three different manufacturing processes. While short, double convergent grinding grooves are created during plunge grinding, grooves in belt grinding are closed, evenly deep and circumferential. Superfinishing results in crossed surface structures that are very alike to those of honing.

Fig. 8 - Surface structures of the different manufactured shaft counterfaces TEST COMPONENTS

Tests were carried out with the reference oil FVA 3, whose parameters are all given by Forschungsvereinigung Antriebstechnik e.V. (FVA) (2007). Table 1 shows the main characteristics of the oil. A particular advantage is that the temperature dependence of the dynamic viscosity η of the oil is also known, as this is required for determining the duty parameter. The given values of the dynamic viscosity η were compared with self-measured values.

Table 2 - Oil parameters FVA 3 (Forschungsvereinigung Antriebstechnik e.V. (FVA), 2007)

Density [kg/m³] 875 – 885 Kinematic Viscosity at 40 °C [mm²/s] 90 – 100

Kinematic Viscosity at 100 °C [mm²/s] 10.2 – 11.2 Viscosity Group (VG) 100

Thermal conductivity [W/(m∙K)] (Schuler, 2014) 0.13

Radial shaft seals made of nitrile butadiene rubber (NBR) of form A were used as sealing rings. Three belt ground and superfinished shaft counterfaces each were tested, as well as a plunge ground shaft counterface, which serves as a reference. Shaft diameter to be sealed was 80 mm.

FRICTION TORQUE MEASUREMENT

Friction torque measurements were carried out on the test rig shown in Figure 9. To measure the friction torque, caused by a radial shaft sealing system, the test chamber together with the radial shaft seal is mounted aerostatically and thus almost frictionless. In this way, the friction torque caused by a sealing system can be determined by measuring the force required to support the test chamber as well as the lever arm of force application. The force is measured via a load cell. The lubrication of the sealing contact is achieved by splashing the sealing contact with oil (V̇ = 800 ml/min) and an open oil circuit tempered to 60 °C. Advantage of the oil splashing in comparison to an oil sump is, that no drag torques through the rotation of the shaft in the oil sump occur. All tests were performed at atmospheric pressure. Before the actual test run began,

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each sealing system was run in for 20 hours at a speed of n = 1194 rpm, which corresponds to a circumferential speed of u = 5 m/s, see Figure 10.

The run-in creates constant conditions for the subsequent test run. During the test collective, 30 speed levels, starting with a speed of n=1 rpm up to n=2500 rpm, were carried out, each lasting one hour. Since the permissible load limits of NBR material are reached at a speed of about n=2200 rpm (Freudenberg Sealing Technologies GmbH, 2015; Trelleborg Sealing Solutions, 2011), the collective was carried out from low to high speeds to avoid a thermal damage of the sealing edge already at the beginning of a test run. The measured values of friction torque and friction coefficients presented in this paper are based on the average values of the last five minutes per speed level. For statistical reason three clockwise and at least two counter clockwise tests were performed with each shaft counterface. The in the following presented results show the average values of all repetitions in one rotational direction, unless explicitly described otherwise.

Fig. 10 - Speed collective of friction test run The sealing gap temperature required for the calculation of the duty parameter or its temperature-dependent components (dynamic viscosity η and mean surface pressure pm) was estimated using the program InsECT. An estimation of the temperature is necessary because the sealing gap temperature cannot be measured without an intervention in the sealing system. Thus, for example, with an optical measurement only the temperature of the shaft counterface in the near of the sealing gap can be measured. The exact procedure for estimating the

Fig. 9 - Friction test rig

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temperature and the effects on individual components of duty parameter diagram are explained in more detail in the following section.

RESULTS

In the following section is first explained how the temperature dependence of the dynamic viscosity η and the radial force was considered in the evaluation. In addition, it is explained how the temperature estimation was performed with InsECT. Afterwards, the results of the analyses are presented in form of friction torque curves and duty parameter diagrams.

Temperature Dependent Parameters and Temperature Estimation

The dynamic viscosity η of the oil was measured on a rheometer up to a temperature of 80 °C. The course of the measured values over the temperature is shown in Figure 11 on the left. In order to be able to calculate the dynamic viscosity η, which is required for the calculation of the duty parameter at sealing gap temperature, the measured course of the dynamic viscosity η was approximated by the Ubbelohde-Walther equation.

To determine the temperature dependence of the radial force, the radial force of three radial shaft seals was measured at five temperature levels from 20 °C to 120 °C. The measurement was carried out in each case after the radial shaft seals had been stored for 48 hours on a shaft at corresponding temperature. The mean values of these measurements are shown in Figure 11 on the right. To approximate the radial force at the sealing gap temperature, a quadratic function was fitted through the measuring points. Before and after each test, the radial force of the radial shaft seals used was measured after they had also been stored for 48 hours on a shaft at room temperature (20 °C). To obtain the radial force at the sealing gap temperature, the approximated function was shifted by the offset between the mean value of the two measured radial forces before and after the test and the mean value of the radial forces of the radial shaft seals used to determine the function. The radial force, which was determined in this way, acting at the sealing gap temperature was used to calculate the duty parameter and the coefficient of friction.

Fig. 11 - Temperature dependence of dynamic viscosity (left) and radial force (right)

The sealing gap temperature for determining the dynamic viscosity η and the radial force in operation was estimated using the open source software InsECT. For temperature estimation using InsECT, the physically motivated ExACT method (Feldmeth et al., 2016) was used. This

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method requires for a temperature estimation the amount of energy, which is introduced into the system by friction. For this we used the frictional power measured in the test runs. Furthermore, for the temperature estimation, some characteristic data of the oil are required, such as kinematic viscosity at two temperatures, density as well as thermal conductivity (see Table 1). The operating conditions are represented by the shaft speed and shaft diameter, the relative oil level, the oil sump temperature, the thermal conductivity of the shaft and the outside air temperature. Due to oil splashing, the relative oil level was assumed to be 0 %, with an oil sump temperature of 60 °C, which stands for the oil temperature. The thermal conductivity of the used shafts made of 100Cr6 is 45 W/(m∙K) (Schaeffler Technologies GmbH & Co. KG, 2014). Figure 12 shows a comparison of the estimated sealing gap temperature (by InsECT) and the measured temperature (by thermo camera) of the shaft counterface. After a speed of about 100 rpm is reached, the sealing gap temperature increases almost linearly. The deviation of about 10 K between both curves for higher speeds can be explained with the measurement of the temperature by thermo camera. Using this method only allows to measure the temperature at the airside of the sealing contact. Based on experience, the actual over temperature in the sealing gap is higher than the temperature value measured on the airside of the sealing contact. This relationship was examined in detail by Feldmeth et al. (2013), (2015). The temperature deviations were elaborately adjusted there using measurements, estimates and CHT simulations. Due to the good agreement of the curves, the estimated values can therefore be regarded as valid.

Fig. 12 - Comparison of estimated and measured temperature in a test run with shaft counterface BG2

Friction Torque – Speed Relation

In Figures 13 to 16, the measured friction torque over the entire speed range analysed is shown on the left-hand side. In order to be able to see the course at low speeds better, the figures on the right-hand side show the measured friction torque exclusively over a speed up to n = 500 rpm. Figure 13 shows, as an example for belt ground shaft counterfaces, three friction torque curves measured clockwise with shaft BG2 and their mean value, which is used in the following for comparison with the other shafts. By analogy, Figure 14 shows individual friction torque

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curves and their mean value when using the superfinished shaft counterface S2. The friction torque curves of individual measurements show a scattering typical for tribological tests. This scatter between the individual measurements is to be judged as low for all tested shafts.

Fig. 13 - Friction torque curves of individual tests and their mean value (shaft counterface BG2)

Fig. 14 - Friction torque curves of individual tests and their mean value (shaft counterface S2)

Figure 15 shows a comparison of the friction torque curves of various belt ground shaft counterfaces with the plunge ground reference as an example of a clockwise direction of rotation. It is noticeable that the friction torque caused increases with a decreasing of the roughness of shaft counterface. Up to a speed of n = 200 rpm, the lowest friction torque is generated with the reference shaft counterface. At higher speeds, the friction torque generated with shaft counterface BG1 is slightly lower. All friction torque curves show a minimum in the speed range of n ≤ 25 rpm. Above a certain speed, which varies slightly depending on shaft conuterface, the friction torque generated by sealing systems remains almost constant.

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Fig. 15 - Mean friction torque curves of belt and plunge (reference) ground shaft counterfaces Friction torque curves of sealing systems with superfinished shafts are shown in Figure 16. As in the case of sealing systems with belt ground shafts, the friction torque generated increases as the shaft counterface becomes smoother. A comparison with the reference system shows that use of a superfinished shaft counterface causes a higher friction torque. The friction torque curve of the sealing system with shaft counterface S3 is the only one that does not show a minimum. In contrast to sealing systems with belt and plunge ground shaft counterfaces, the friction torque caused by sealing systems with superfinished shaft counterfaces decreases continuously with increasing speed after reaching a maximum.

Fig. 16 - Mean friction torque curves of superfinished and plunge ground (reference) shaft counterfaces Friction Torque – Duty Parameter Relation In analogy to the procedure with the friction torque curves, the curves of the individual tests and their mean values are shown in the duty parameter diagram in Figure 17 and Figure 18 for the shaft counterfaces BG2 and S2. In both figures, the curves for a rotation in clockwise direction are shown on the left and in counter clockwise direction on the right. The scattering occurring between repeated tests is small.

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Fig. 17 - Duty parameter diagram of individual tests and their mean value (shaft counterface BG2;

rotational direction left cw and right ccw)

Fig. 18 - Duty parameter diagram of individual tests and their mean value (shaft counterface S2; rotational

direction left cw and right ccw)

The duty parameter diagrams of sealing systems with belt ground shaft counterfaces are shown in Figure 19 and those of sealing systems with superfinished shaft counterfaces in Figure 20. With the exception of the smoothest superfinished shaft counterface (S3), all graphs in the duty parameter diagram show a typical course with an area of liquid friction as well as an area of mixed friction and a transition point (X in diagrams) between both. In the case of belt ground shaft counterfaces and a clockwise rotation, it can be seen that with a decreasing surface roughness (shaft BG1 to BG3) a higher coefficient of friction is caused (Figure 19, left side). Contrary, with a counter clockwise rotation the roughest shaft counterface caused the highest coefficient of friction (Figure 19, right side). The dependence of the curves in the duty parameter diagram on the rotational direction of the shaft counterface can be explained either by the shaft counterface itself, e.g. micro lead structures, or by the elastomeric lip seal. Further

transition point transition point

transition point transition point

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investigations are currently carried out to prove the influence of the elastomeric lip seal on a rotational direction dependent behaviour.

Fig. 19 - Influence of surface roughness of belt ground shaft counterfaces on the tribological condition of an elastomeric lip seal (rotational direction left cw and right ccw)

Superfinished shaft counterfaces also cause with a decreasing roughness (shaft S1 to S3) a higher coefficient of friction (Figure 20). Furthermore, with a decreasing roughness the transition point is moving to smaller values of the duty parameter. This means that with smoother shaft counterfaces the area of liquid friction is already reached at smaller speeds. This however, caused a higher friction in total. In the test runs, the transition point is at latest reached at a circumferential speed of about 0.1 m/s. Tests with superfinished shaft counterfaces lead, in case of the smoothest shaft, to a course in the duty parameter diagram without a transition point. Thus, it can be supposed, that in this test runs from beginning liquid friction and not mixed friction occurred. Due to leakage during the test runs in counter-clockwise rotational direction, the sealing systems with the superfinished shafts S2 and S3 show, dependent on the rotational direction of the shaft counterface, different courses in the duty parameter diagram. Responsible for the leakage can be the measured left-hand macrolead of these shaft counterfaces. In the remaining test runs no leakage occurred.

Fig. 20 - Influence of surface roughness of superfinished shaft counterfaces on the tribological condition of an elastomeric lip seal (rotational direction left cw and right ccw)

transition point transition point

transition point transition point

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Above a value of about G = 5 -7 10-7 the friction of sealing systems with a superfinished shaft counterface decreases rapidly (red marking in Figure 20). This decrease is according to results of Schuler (2014) and Ott (1983) due to a high loss of viscosity and radial force in consequence of a high heat impact. Contrary, with belt and plunge ground shafts above a value of about G = 10-6 the friction increases strongly (blue marking in Figure 19).

In comparison with the plunge ground reference shaft counterface, each sealing system with an alternatively manufactured shaft counterface, with exception of belt ground shaft BG1, cause in clockwise rotation higher coefficients of friction. When turning counter clockwise, the curves of all sealing systems in the duty parameter diagram are in a smaller range of the coefficient of friction.

SUMMARY AND CONCLUSIONS

In sealing systems with belt ground shaft counterfaces, with exception of the roughest shaft counterface, a higher friction torque and a higher coefficient of friction occurs than in the reference sealing system with a plunge ground shaft counterface. The measured friction torque remains almost constant after reaching a maximum value. As the shaft counterface becomes smoother, friction (friction torque and coefficient of friction) increases. All duty parameter diagrams show a minimum, i.e. in all sealing systems with belt ground shaft counterfaces, mixed or liquid friction occurs, depending on the operating conditions. The transition point from mixed to liquid friction occurs with belt ground shafts at slightly lower duty parameters than with the reference shaft.

The use of superfinished shafts also results in a higher friction torque and a higher coefficient of friction than with the reference system. Similar to belt ground shaft counterfaces, friction (friction torque and coefficient) increases with decreasing roughness of the shaft counterface. The smoother the shaft counterface is, the less pronounced minimum friction is and the further the transition point shifts to a lower duty parameter. This goes so far that no minimum occurs with the smoothest shaft examined. From this it can be concluded that pure fluid friction occurs in the sealing contact of very smooth superfinished shaft counterfaces already at speeds of n = 1 rpm. This refuted the hypothesis mentioned at the beginning, for the analysed shaft counterfaces, that the transition point shifts to higher operating parameters with decreasing roughness of the shaft counterface. Contrary, to the use of belt ground shafts, with superfinished shafts and an increasing duty parameter the emerged friction torque continuously decreases after a maximum was reached.

In Figure 21 are depicted the scatter ranges of the duty parameter diagrams of the three differently manufactured types (belt and plunge grinding, superfinishing) of shaft counterfaces in a clockwise direction of rotation. The curves show the mean values of all test runs of one manufacturing method and the scatter range indicates the maximal and minimal friction torque of all these test runs. The diagram clearly shows that at low duty parameters (G less than 10-7) with belt ground shaft counterfaces manufactured to the appropriate setting values, less friction is generated than in the two reference sealing systems with a plunge ground shaft counterface. Contrary, the scattering range of the friction coefficient from sealing systems with plunge ground shaft counterfaces is narrower than with belt ground shaft counterfaces. This means that when using plunge ground shaft counterfaces, the resulting friction can be negatively influenced to a lesser extent by manufacturing parameters. The diagram also makes it clear that higher coefficients of friction occur when superfinished shaft counterfaces are used, almost independently of the operating point (the duty parameter).

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Fig. 21 - Scatter ranges of the duty parameter diagrams for the variously manufactured shaft counterfaces

for a clockwise direction of rotation

Based on the results obtained from the tests with superfinished shaft counterfaces, it is obvious that such shaft counterfaces are not a suitable alternative for sealing systems. Contrary, if the shaft counterfaces are manufactured in the right way, belt ground shaft counterfaces seem to be an alternative for operating point with small duty parameter. The caused friction of sealing systems with belt ground shaft counterfaces can be regulated by their roughness. Due to worse tribological conditions in sealing systems with belt ground shaft counterfaces, higher friction losses, an unsteady wear and leakage behaviour for operating points with a higher duty parameter (Schulz et al., 2018a; Schulz et al., 2018b) occur. Thus, belt ground shaft counterfaces cannot be fully recommended for the use as sealing counterfaces.

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[3] Deutsches Institut für Normung e.V.: DIN 3761-1: Radial-Wellendichtringe für Kraftfahrzeuge - Begriffe; Maßbuchstaben; zulässige Abweichungen; Radialkraft, Berlin: Beuth. Januar 1984.

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Track-F

Tribology and Surface Engineering

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