influence of different shaft surface finishes on the

The 11th International Fluid Power Conference, 11. IFK, March 19-21, 2018, Aachen, Germany

2 The tribological system of radial shaft seals

The tribological system of radial shaft seals consists of the sealing element itself, the fluid, the ambient conditions and the shaft counterface, Figure 1. The mounting situation and the operating conditions have to be considered as well. Interactions between the main elements of the tribological system can lead to a critical operating condition of the sealing system.

Figure 1: Tribological system of radial shaft seal /5/.

Whereas the sealing element itself and the fluid are bought-in-parts, the shaft counterface usually is manufactured internally or by a supplier according to self-defined specifications. Because of that, the shaft counterface is one of the main sources of failures. During the manufacturing process already, small changes can have a negative impact on the quality of the shaft counterface. And despite of this critical influence, the manufacturing parameters are often undefined /5/. If the standard shaft manufacturing method plunge grinding is replaced by an alternative man-ufacturing method like belt grinding or superfinishing, there is always a risk of negative influences on the tribo-logical system of seals.

2.1 Sealing mechanisms of radial shaft seals

In the dynamic operating condition, the elastomeric lip seal forms an active sealing mechanism, which prevents leakage of the sealed fluid. The sealing mechanism continuously pumps fluid from the airside to the oilside. This sealing mechanism can be explained by different so-called pumping hypotheses. The three main hypotheses for the fluid pumping mechanism all base on an asymmetric contact pressure distribution. The left side of Figure 2 shows the pumping hypothesis, which relies on axial wear structures and was developed by Kammüller /6/. Due to the rotation of the shaft and the asymmetric contact pressure distribution the wear structures are distorted in circumferential direction. Thus, the wear structures form valleys and ridges. As a result, the fluid entrained by the shaft is axial redirected by the wear structures and a dynamic pressure is built up in front of the zone with maximum contact pressure. If the dynamic pressure reaches a value above the average contact pressure of the seal, the seal lip lifts off and the fluid is conveyed across the maximum pressure contact zone. Because of the asymmetric pres-sure distribution, on the airside of the sealing lip the wear structures are on a wider area in contact with the shaft surface than on the oilside. Consequently, the pumping volume from the air- to the oilside is bigger than reverse.

Figure 2: Pumping hypothesis with axial wear structures (left) and microasperities (right) /7/.


Influence of different shaft surface finishes on the tribological and functional be-haviour of radial shaft seals

Markus Schulz*, Matthias Baumann*, Frank Bauer* and Werner Haas*

University of Stuttgart, Institute of Machine Components (IMA), Pfaffenwaldring 9, D-70569 Stuttgart, Ger-many*

E-Mail: [email protected]

The shaft counterface of a radial shaft seals is usually plunge ground. With the aim to reduce costs and production times, many companies try to use new and alternative manufacturing processes for that task. For example, belt grinding and superfinishing methods are frequently considered and used. Result of these often-unreflecting changes of the manufacturing methods, is often leakage and increased wear of the sealing components as well as other related problems. Because, there is only little information about the functional behaviour of this types of surface finishes in terms of sealing applications, an experimental investigation has been carried out. This paper presents the results of these experiments.

Keywords: Radial Shaft Seals, Surface Topography, Failure Analysis, Shaft Manufacturing Methods Target audience: Mechanical Engineering, Manufacturing Technology, Automotive Technology

1 Introduction

Elastomeric radial shaft seals are the most commonly used seals throughout the industry. At the gearbox output of cars, for example, they are applied a million times a year. A radial shaft sealing is a tribological system and consist of several components: the elastomeric lip seal itself, the shaft counterface as well as the fluid, which is to seal. Therefore, the surface finish of a shaft counterface influences the functional and tribological behaviour of a sealing system to a high degree.

Plunge grinding is the standard manufacturing method for shaft sealing counterfaces. To reduce costs and produc-tion times, alternative, faster and cheaper manufacturing techniques are more frequently considered and used. Belt grinding and superfinishing are examples for such alternative manufacturing methods. While plunge grinding has been the focus of research during the last 30 years, these newer and cheaper alternative manufacturing methods have been less investigated in terms of sealing applications. Furthermore, tolerance limits of surface roughness parameters, which guarantee a secure functioning of sealing systems, are still unknown for these manufacturing methods. Because the roughness specifications, established according to the current standards, are only valid for plunge ground shafts and cannot be applied to alternative surface finishes.

According to DIN 3760 and DIN 3761 as well as information from seal manufacturers, a shaft sealing counterface has to be free of lead /1–3/. To achieve this requirement using plunge grinding, the sanding disc has to be well-dressed and long spark-out times have to be applied /3/. Due to the kinematics of the manufacturing process, belt ground surfaces should be free of lead /4/. Therefore, at first glance, belt ground shafts seem to be well suited as sealing counterfaces. Nevertheless, with belt ground and superfinished counterfaces problems with leakage or increased wear often occur. Hence, in this paper is analysed the influence of belt grinding and superfinishing methods on the surface topography of sealing counterfaces and consequently on the tribological and functional behaviour of radial shaft seals.

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3 Experimental procedure

In this chapter follows an explanation of the different manufacturing methods, which have been used to manufac-ture the shaft surface finishes. Afterwards the analysis and measurement methods as well as the test components are introduced.

3.1 Manufacturing of the shaft counterfaces

Hardened sleeves (100Cr6) were belt ground or superfinished with the superfinish-add-on device depicted in Fig-ure 4. The superfinish-add-on device was mounted at the tool holder to the lathe. The grinding belt was pressed against the shaft counterface by the pressure roller and driven by the under role. For a well-functioning manufac-turing process, the contact force can be specified in a range up to maximal 280 N. The grinding belt is not axially fixed. Thus, problems with lead can occur. By a feed rate of the belt it is ensured that always new, fresh abrasive grits are in contact with the shaft counterface. For the manufacturing of superfinished shaft counterfaces the pres-sure roller can be reciprocated axially along the shaft surface. To obtain counterfaces with different roughness, grinding belts with different grain sizes (180, 360, 500, 800 and 1200) were used. The manufacturing of the coun-terfaces was carried out by a stepwise machining of each shaft counterface with grinding belts with different grain sizes. After each manufacturing step, the required number of counterfaces was taken out of the further manufac-turing process.

Figure 4: Superfinish-add-on device Supfina 210 /4/

3.2 Surface analysis of the shaft counterfaces

Surface analysis of the shaft counterfaces contained measurements of the roughness and lead. The lead analysis was classified according to the lead categories in macro- and microlead measurements. Macrolead was analysed by the CARMEN-method /12/ and microlead by the IMA-Microlead® Analysis /7, 13/, which was developed at the Institute of Machine Components of the University of Stuttgart.

The CARMEN-method allows the measurement of circumferential periodical macrolead and is based on 72 tactile measurement traces, which are measured parallel to the axis and are evenly distributed over the circumference (or a 36° section) in a distance of 5° (or 0,5°). These measurement data are used to determine the lead-parameters (lead angle, lead depth, period length, lead thread, …) according to the Daimler Werksnorm MBN31007-7 /12/. In Figure 6 are depicted three-dimensional so-called pseudo-surface topographies, which are composed from the single two-dimensional measurement traces over 36° and 360° of a belt ground shaft counterface. Whereas the pseudo-surface topography of the 36° section shows circumferential and closed structures, the pseudo-surface topography of the whole circumference (360°) shows, that the structures differ from the circumferential direction. Thus, over the whole circumference exist sections of right-hand (no. 1) and left-hand (no. 2) lead as well as sec-tions, which are free of lead (no. 3).


The right side of Figure 2 shows a second sealing hypotheses, which was developed by Müller and Kammüller /8/. They found, that an active sealing mechanism also exists, when the sealing edge exhibits no axially directed wear structures. Therefore, they postulated a further hypothesis for the fluid pumping mechanism of seals that emanates from stochastically distributed microasperities on the sealing edge. Due to the shaft movement, there are micro-scopic pressure fields in front and behind of the contact zone of the microasperities and the shaft surface. The pressure in front of the microasperities is higher and behind them smaller than the mean value of the local pressure. The drag flow is thus axial redirected or aspirated and can cross the zone of maximum contact pressure (pressure equator) due to a higher local pressure in both directions. Because of a flatter pressure gradient on the airside of the sealing edge, follows from a statistical perspective, that more fluid cross the maximum contact pressure zone from the air- to the oilside. On the border of the sealing edge fluid is aspirated into the sealing gap by cavitation zones /9, 10/.

As third hypotheses for the fluid pumping mechanism of seals was postulated the so-called oscillation theory by Jenisch /11/. This theory expects, that the sealing edge of an elastomeric lip seal, which is tilted or eccentrically mounted in relation to the shaft axis, executes an axial oscillating motion. Due to this swipe movement and the asymmetric contact pressure distribution a fluid transport from the air- to the oilside emerges.

All mentioned sealing hypotheses describe different pumping mechanisms, which forms in superposition the active sealing mechanism of radial shaft seals. In a well-designed sealing system, there is an equilibrium between fluid, which enters the sealing gap, and fluid, which is pumped back from the elastomeric lip seal. Consequently, the equilibrium leads to a hydrodynamic lubricated sealing contact, which results in good tribological conditions and simultaneously prevents leakage. The fluid transport dissipates the frictional heat, too. Due to these reasons, the active sealing mechanism is crucial for the function of a shaft seal system.

2.2 Influence of shaft lead on the sealing function

The shaft counterfaces can contain structures (e.g. lead), that also pump fluid. In Figure 3 is depicted, that the rotation of the shaft can lead to dry run and increased wear, if the transport direction of the shaft points into the sealing system, or to leakage, if the transport direction of the shaft points in the direction of the airside. Further-more, Figure 3 shows the differences between circumferential macrolead and microlead as well as that the appear-ance of leakage or dry run is dependent on the rotational direction of the shaft.

Figure 3: Influence of lead on the function of the tribological shaft seal system /4/

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Figure 6: Comparison of the angle- and volume distribution curves of both plunge ground (shaft Ref 1 (above) and shaft Ref 2 (middle)) and a belt ground (below) counterfaces.

Table 2 shows the roughness parameters of the investigated shaft counterfaces. The roughness parameters of the belt ground shaft counterfaces are with exception of shaft counterface BG 1 within the limits given by standard DIN 3760 (Ra = 0.2 – 0.8 µm, Rz = 1 – 5 µm and Rmax = 6.3 µm) /1/. However, the roughness parameters of the superfinished counterfaces were with exception of S 1 too small and thus out of range of the standardized rough-ness parameters.

Shaft counterface BG 1 BG 2 BG 3 BG 4 BG 5 S 1 S 2 S 3 S 4 Ref 1 Ref 2

Ra [µm] 1.0 0.66 0.44 0.31 0.28 0.29 0.13 0.08 0.03 0.32 0.19

Rz [µm] 6.81 4.39 3.29 2.47 2.16 2.47 1.25 0.79 0.40 2.75 1.61

Rmax [µm] 8.0 5.57 3.84 3.1 2.66 3.04 1.57 1.01 1.61 3.26 2.01

Table 2: Roughness parameters of the used counterfaces.


Figure 5: Pseudo-surface topography of a belt ground shaft counterface over 36° (left) and 360° (right)

If the macrolead analysis however shows no lead thread, the lead is called zero-lead. Reason for the axial deviations can be an axial tolerance of the grinding belt or an inclined respectively eccentrical clamped shaft counterface. The lead parameters of the measured counterfaces are shown in Table 1. All belt ground counterfaces and the reference shafts were free of macrolead. Contrary, the superfinished counterfaces show macrolead, which depend-ent of the shaft manufacturing is more or less distinct.

Shaft counterface BG 1 BG 2 BG 3 BG 4 BG 5 S 1 S 2 S 3 S 4 Ref 1 Ref 2

Lead angle [°] 0 0 0 0 0 0.18 0.40 -0.13 0.04 0 0

Lead depth [µm] 1.13 1.14 0.41 0.39 0.32 0.08 0.02 0.02 0.01 0.08 0.2

Period length [mm] 0.21 0.35 0.24 0.25 0.27 0.20 0.35 0.29 0.16 0.27 0.11

Lead thread - - - - - 4 5 -2 1 - -

Table 1: Macrolead parameters of the measured counterfaces.

The IMA-Microlead® analysis is a computational procedure, which base on optical measurement data. To obtain these optical measurement data, surface topographies of the shaft counterfaces are needed. These surface topogra-phies are obtained by a white light interferometer. The measurement data are filtered, separated and statistically analysed. Following this, the size, position, angular orientation and the volume of a single grinding groove can be determined. Outgoing of these measured quantities an angle- and a volume distribution curve are generated. Sur-faces without microlead show a symmetric distribution of the left and right oriented structures with a maximum in the direction of circumference (at 0°). In Figure 6 are depicted the angle- and volume distribution curves of both plunge ground reference counterfaces and a belt ground shaft counterface. The distribution curves of all shaft counterfaces, which were belt ground, are narrower than those of a plunge ground shaft counterface. In case of the second plunge ground shaft counterface, the distribution curves are asymmetric and therefore show a right-hand microlead. On the contrary, the angle- and volume distribution curves of all remaining counterfaces have a maxi-mum at 0° or 0.1° and are symmetric. Consequently, these counterfaces are free of microlead. The IMA-Micro-lead® calculation algorithm is not capable of detecting crossed structures correctly. This topic is under current work.

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Figure 8: Endurance test rig /4/.

The test runs were performed for a duration of 96 hours with an oil temperature of 60 °C and a circumferential speed of 5 m/s. Furthermore, the test chambers were vented, because the test should be run under atmospheric pressure. The choice of these test parameters was according to the public funded project Lastkollektive I /15/.

In order to evaluate the test runs, leakage and wear measurements as well as a visual analysis of the worn sealing edges were conducted. To measure the amount of leakage, the leaking oil is collected in a beaker glass and then weighed. The wear widths of the sealing edges were measured with a digital incident light microscope and a device for the correct alignment of the sealing edges. This alignment is necessary, due to meet the requirement of a per-pendicular view on the sealing edge. Therefore, an inaccurate measurement of the wear widths is excluded. All wear widths shown in section 4.2 are averaged values of 16 single measurements of the worn contact area of the sealing edges (see Figure 12). The measurement position were hereby evenly distributed around the circumference of the seals.

3.4.2 Friction test rig

The test rig, which has been used for the frictional torque measurements, is depicted in Figure 9. Its main compo-nents are the high-speed motor spindle as well as the aerostatic, frictionless bearing of the test chamber. As a result of the aerostatic bearing the friction force can be measured with a force sensor. Consequently, the friction torque can be calculated from the measured friction force and the lever arm of the point of load application. The oil was tempered by an extern-tempering unit. By doing so, it was possible to measure the friction torque of the shaft sealing without other falsifying effects such as the influence of electrical connections of the heating elements.

Figure 9: Friction test rig.

The friction measurements were performed with an oil temperature of 60 °C and an oil level, which reached to the shaft axis. Furthermore, the tests were run under atmospheric pressure. Table 4 shows the test procedure of the friction test runs. In order to achieve similar test conditions over the whole duration of the test run, bevor each test run the sealing system was running-in for an hour. Hence, the running-in wear of the sealing system have a minimal


Figure 7 shows the significant differences in the shaft surface topographies caused by the various manufacture methods. Plunge grinding causes, for example, individual short, non-closed and double-convergent grinding grooves, whereas belt grinding results in uniformly deep, circumferential grooves. The so-called superfinishing processes cause crossed grooves, which are much alike to structures at honed surfaces.

Figure 7: Topographies of the three different shaft surfaces.

3.3 Test components

The tests were run with the high temperature resisting, fully synthetic engine oil Fuchs Titan Supersyn Longlife SAE 0W-30. In order to allow a comparison of the different test results, all tests were run with an oil level, which reached to the shaft axis. Table 3 summarizes the most important parameters of the oil.

Viscosity class SAE 0W-30

Density [kg/m³] 843

Kin. Viscosity at 40 °C [mm²/s] 67

Table 3: Oil parameters /14/.

The deployed seals made of nitril butadien rubber (NBR) were of the form A according to the standards DIN 3760 and DIN 3761 /1, 2/. Two plunge ground shaft counterfaces were appointed as reference for the alternatively manufactured shaft counterfaces. The shaft counterface Ref 1 exhibited, as already mentioned above, microlead, whereas reference Ref 2 was free of microlead. The diameter of the analysed counterfaces was 80 mm.

3.4 Test runs

Different test runs were applied to obtain comprehensive data about the tribological and functional behaviour of the analysed sealing system. Firstly, endurance tests were performed to achieve data about the functioning of the sealing system and the wear of the elastomeric sealing lip as well as the shaft counterface. Secondly, friction tests were conducted to achieve information about the influence of the different surface structures on the lubrication condition of the sealing system.

3.4.1 Endurance test rig

The endurance tests with the analysis of the sealing edges and the leakage are the basis of this study. Figure 8 shows the deployed endurance test rig, which consists of six modules with in each case two test chambers. Since one spindle is driving the shaft counterfaces of a module, the rotational direction of the shafts of both test chambers is always converse. The shaft counterfaces are mounted to the test rig with a HSK-interface and a HSK-collet chuck.

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causes a higher average wear width. The rotational direction of the plunge ground shaft with microlead (Ref 1) was counter clockwise. Therefore, the shaft transported the oil in the sealing gap, according to Figure 3, back to the oilside. Consequently, the sealing edge was only fractionally lubricated and as a result, increased wear oc-curred. Due to this, the wear of the sealing lip is already to be assessed as critical. Hence, for a secure functioning of the sealing system the wear widths of the elastomeric lip seals with belt ground and superfinished shaft coun-terfaces were compared to the plunge ground shaft counterface without microlead (Ref 2). The roughest belt ground shaft counterfaces (BG 1 and BG 2) induce even more wear than the plunge ground shaft counterface with microlead (Ref 1). Only the belt ground shaft counterface BG 3 leads to a similar wear width as the reference shaft counterface. The average wear widths of the superfinished shaft counterfaces are with exception of the roughest shaft counterface (S 1) smaller than the reference wear width (Ref 2). Around the circumference, the wear widths of all elastomeric lip seals varied.

Figure 11: Wear widths of the elastomeric lip seals.

These wear widths allow only limited conclusions on the wear behaviour, since the sealing edges, which were used with a belt ground shaft counterface, partially exhibited strong hardenings and on the other hand the sealing edges, which were in use with a superfinished shaft counterface, were well lubricated due to leakage. These hardenings, depicted in Figure 12, result in a slower increase or a constant value of the wear width. In the area of the hardening arose circumferential grooves and axial cracks. The combination of hardening, grooves and cracks lead in a longer period of use to an unexpected and sudden failure of the sealing system. Furthermore, also in the area of non-hardened elastomer are recognisable areas of different lubrication conditions. Areas with an insufficient lubrication are strongly coloured because of higher temperatures.

Figure 12: Hardening of a sealing edge after use with a belt ground counterface.

4.2.2 Counterface

Figure 13 shows a running track of a seal on the smoothest and roughest superfinished (S 4 and S 1) shaft coun-terfaces. On the smoothest counterface the most of the crossed surface structures were worn in the area of the elastomeric lip seal. Similarly, on the roughest counterface fine structures were worn, but broader and deeper structures were still remained. Due to the wear of the structures, the leakage behaviour described in section 0 can


influence on the measurement of the friction torque. Following the running-in phase, the first step of the test run was conducted under the same test conditions as the running-in. The results shown in section 4.3 represent the average value of the last five minutes of the respective step of the circumferential speed. In contrast to the endur-ance tests, friction measurements were only conducted with the belt ground and plunge ground reference shaft counterfaces with microlead.

Circumferential speed Duration [h]

Running-in 5 m/s 1

Test run 9 steps from 0.1 m/s to 15 m/s 0,5

Table 4: Test procedure of the friction test runs.

4 Results

In this chapter are presented the test results of the endurance and friction tests. Firstly, the results of the leakage analysis are shown. Afterwards, the wear of the elastomeric lip seals and the shaft counterfaces is investigated. Subsequently, are illustrated the results of the friction test runs. At last follows an interpretation of all results.

4.1 Leakage

In all test runs with belt ground shaft counterfaces no leakage occurred. In comparison, the test runs with super-finished counterfaces all showed more or less leakage. Figure 10 depicts the measurable leakage of test runs with superfinished shaft counterfaces over the test duration. The most leakage occurred in the first 24 hours of the test runs.While the leakage of the test run with the smoothest superfinished shaft counterface (S 4) continuously de-creased, in case of the second depicted test run, the leakage decreased rapidly after the first 24 hours of test dura-tion. Both test runs (counterfaces S 1 and S 3), which are not shown in Figure 10, had an oil meniscus at the airside of the sealing lip, however the volume of the leakage was too small for a measurement. The reference sealing system with a plunge ground shaft counterface was tight.

Figure 10: Leakage of test runs with superfinished shaft counterfaces.

4.2 Wear

The analysis of the wear was separated into the study of the elastomeric lip seals and the shaft counterfaces. How-ever, the main part of the analysis was the study of the elastomeric lip seals.

4.2.1 Elastomeric lip seal

Figure 11 depicts the average as well as the maximal and minimal measured wear widths of each elastomeric lip seal. A comparison of the wear widths of the test runs with both plunge ground shafts shows that the microlead

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Figure 15: Correlation between the friction torque and the arithmetic mean roughness Ra at a circumferential speed of 2.5 m/s

5 Summary and Conclusion

The surface analyses of the belt ground shaft counterfaces illustrate, that it is possible to manufacture belt ground shaft counterfaces, which are free of lead. However, it is not guaranteed, that belt ground shaft counterfaces are free of lead. Furthermore, the lead-free state of a shaft counterface cannot be an exclusive criterion for a suitable sealing counterface.

While superfinished shaft counterfaces lead to a failure of a sealing system due to leakage, belt ground counterfaces were temperature damaged because of an insufficient lubrication. Furthermore, sealing systems with a belt ground counterface cause a higher friction torque than by the use of a plunge ground counterface. The reason for the damage and higher friction is an insufficient replacement of the oil in the sealing gap. Figure 16 shows that the oil replacement is disturbed by the circumferential and closed grooves of a belt ground counterface. Therefore, the airside of the sealing edge is insufficiently lubricated. On the contrary, at a plunge ground counterface the oil replacement leads to a sufficient lubricated sealing edge as well as a heat replacement. On plunge ground counter-faces double convergent grooves, which one can be seen in Figure 16, are responsible for the oil replacement and the lift off of the sealing edge. The difference between the grooves of a belt ground and a plunge ground counter-face is that the depth and width of a double convergent groove vary whereas the depth and width of belt ground grooves remain almost constant. Therefore, double convergent grooves lead to a build-up of a pressure and due to the dragged oil furthermore the sealing edge is lift off and an elasto-hydrodynamic sealing gap is generated. In the case of belt ground counterfaces the oil in contrary is only dragged around the circumference.

Figure 16: Lubrication conditions of a sealing edge /4/.

Superfinished and belt ground shaft counterfaces show either leakage, increased wear or damages due to an insuf-ficient oil replacement in the sealing gap. Therefore, a secure sealing with counterfaces manufactured in this way cannot be guaranteed. Furthermore, the wear of superfinished counterfaces is significant, in spite of the leakage and consequently well lubrication of the sealing edge.


be explained. The fine structures are still present at the beginning of the test runs. Therefore, the structures can transport oil and furthermore cause leakage. This pumping function of the structures declines with increasing wear consequently the leakage also decreases. Are considered the behaviour of leakage and wear, it can be conducted, that fine structures have a greater impact on the fluid transport than broader and deeper ones. This conclusion is founded in the faster decrease of the leakage in the first 24 hours of a test run with the roughest superfinished counterface and the concurrent wear of fine surface structures. In comparison to this, on the plunge and belt ground shaft counterfaces a running track of a seal cannot be detected.

Figure 13: Sealing track on the smoothest (left) and roughest (right) superfinished (S 4 and S 1) counterface.

4.3 Friction torque

The results of the friction torque measurements are illustrated in Figure 14. In order to receive a better overview only the friction torque of the smoothest and roughest belt ground as well as of the plunge ground shaft counterface with microlead are depicted. The friction torque of the remaining belt ground counterfaces (BG 2 to BG 4) were between the shown friction torques. At circumferential speeds up to 1 m/s the sealing systems with the belt ground counterfaces, excepted the smoothest counterface, caused a lower friction torque than the sealing system with the plunge ground counterface. Although the plunge ground counterface cause an increased friction torque due to microlead and an insufficient lubrication, with increasing circumferential speed the caused friction torque of all sealing systems with belt ground counterfaces was higher.

These results can be explained by the surface structures shown in Figure 7. While the plunge ground counterface contains double convergent grooves, i.e. grooves with varying depth and width, the structures of belt ground coun-terfaces are almost constant deep and wide around the circumference. As a result, in sealing systems with a plunge ground counterface within the dragged oil is build up a pressure, which leads to a lift off of the elastomeric lip seal. At low circumferential speeds the dragged oil volume and consequently the pressure build up is insufficient for a lift off of the sealing edge.

Figure 14: Friction torque.

Figure 15 depicts a correlation between the arithmetic mean roughness Ra and the friction torque. In the area of Ra = 0.4 – 0.6 µm the sealing systems with belt ground counterfaces caused a minimal friction torque. The refer-ence sealing system induced in comparison a significant lower friction torque (approx. 17.4 %).

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References

/1/ Deutsches Institut für Normung e.V., DIN 3760, Radial-Wellendichtringe, Berlin. September 1996.

/2/ Deutsches Institut für Normung e.V., DIN 3761-2, Radial-Wellendichtringe für Kraftfahrzeuge - Anwen-dungshinweise, Berlin: Beuth. November 1983.

/3/ Prem, E.; Vogt, R., The Simmering - Basic for preventing damage. Firmenschrift, Freudenberg Simrit GmbH & Co. KG, Weinheim, 2008.

/4/ Schulz, M., Bauer, F., Haas, W., Einfluss supergefinishter / bandgeschliffener Dichtungsgegenlaufflächen auf die Funktion von Radial-Wellendichtungen. 19th International Sealing Conference (ISC), Stuttgart, 12.-13. Oktober 2016; Fluidtechnik; Frankfurt am Main: Fachverband Fluidtechnik im VDMA e.V, 2016, S. 228–245 - ISBN 978-3-8163-0684-9.

/5/ Baumann, M., Bauer, F., Haas, W., Messung, Analyse und Bewertung von Dichtungsgegenlaufflächen für das Tribo-System Radial-Wellendich¬tung. 18th International Sealing Conference (ISC), Stuttgart, Ger-many, October 8-9, 2014; Frankfurt am Main: VDMA Fluidtechnik, 2014, S. 627–639 - ISBN 978-3-00-046879-7.

/6/ Kammüller, M., Zum Abdichtverhalten von Radialwellendichtringen. Dissertation, 1986, Universität Stuttgart, Institut für Maschinenelemente.

/7/ Baumann, M., Abdichtung drallbehafteter Dichtungsgegenlaufflächen - Messung, Analyse, Bewertung und Grenzen. Dissertation, 2017, Universität Stuttgart, Institut für Maschinenelemente.

/8/ Kammüller, M., Müller, H. K., Physikalische Ursachen der Dichtwirkung von Radialwellendichtringen. Physical Aspects of the Sealing Mechanism of Elastomer Rotary Shaft Seals. ATZ Automobiltechnische Zeitschrift, Springer Verlag. 1986, 88 (1), S. 39–45.

/9/ Stakenborg, M. J. L., On the Sealing and Lubrication Mechanism of Radial Lip Seals. Dissertation, 1988, University of Eindhoven.

/10/ Nakamura, K., Kawahara, Y., An Investigation of Sealing Properties of Lip Seals Through Observations of Sealing Surfaces Under Dynamic Condition. 10th International Conference on Fluid Sealing, Inns-bruck, Austria, 03.-05. April, 1984; Bedford: British Hydromechanics Research Association, 1984, S. 87–105 - ISBN 9780906085899.

/11/ Jenisch, B., Abdichten mit Radial-Wellendichtringen aus Elastomer und Polytetrafluorethylen. Disserta-tion, 1991, Universität Stuttgart, Institut für Maschinenelemente.

/12/ Mercedes-Benz, Norm MBN 31007-7, Geometrische Produktspezifikationen (GPS) - Oberflächenbe-schaffenheit Mess- und Auswerteverfahren zur Bewertung von drallreduzierten dynamischen Dichtflä-chen. September 2008.

/13/ Baitinger, G., Multiskalenansatz mit Mikrostrukturanalyse zur Drallbeurteilung von Dichtungsgegenlauf-flächen. Dissertation, 2011, Universität Stuttgart, Institut für Maschinenelemente.

/14/ Fuchs Schmierstoffe GmbH, TITAN Supersyn Longlife SAE 0W-30, URL: https://www.motoroel.com/me-dia/pdf/58/ae/94/Motoroel-SUPERSYN-LONGLIFE-SAE-0W-30.pdf. [abgerufen am 27.11.2017].

/15/ Eipper, A., Bauer, F., Haas, W., Lastkollektive RWDR. Gestaltung von Lastkollektiven zur Prüfung von Radial-Wellendichtringen. Abschlussbericht FVA Vorhaben Nr. 696, IGF-Nr. 17580 N, Frankfurt am Main: FVA, 2016.

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