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Tribologie und Schmierungstechnik EDITOR IN CHIEF MANFRED JUNGK 5-6 _ 24 VOLUME 71 Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie An Official Journal of Österreichische Tribologische Gesellschaft An Official Journal of Swiss Tribology Issue 5-6 | 2024 Volume 71 Editor in chief: Dr. Manfred Jungk Tel.: +49 (0)6722 500836 eMail: jungk@verlag.expert www.mj-tribology.com Editorial director: Ulrich Sandten-Ma Tel.: +49 (0)7071 97 556 56 / eMail: sandten@verlag.expert Editor: Patrick Sorg Tel.: +49 (0)7071 97 556 57 / eMail: sorg@verlag.expert Dr. rer. nat. Erich Santner Tel.: +49 (0)2289 616136 / eMail: esantner@arcor.de Contributions marked with the author’s initials or full name represent the author’s opinion, not necessarily that of the editorial office. We take no responsibility for unsolicited contributions. The author is responsible for obtaining the rights to pictures. 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ISSN 0724-3472 ISBN 978-3-381-13281-2 Imprint Tribologie und Schmierungstechnik Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie | An Official Journal of Österreichische Tribologische Gesellschaft | An Official Journal of Swiss Tribology Editorial 1 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0032 According to the Agora Energiewende organization carbon dioxide equivalents, including methane and nitrous oxide, fell by three percent to 656 million tons between 2023 and 2024. With this Germany is meeting its own national climate protection targets but is lagging the EU’s target which was missed by 12 million tons. Compared to 1990 the reduction was 48 % with the biggest contribution coming from the energy sector utilizing more wind and solar power. In balance with industry increase by 3 million tons, agriculture being constant, buildings and transport reduced 2 million tons each the total reduction of 18 million tons was mathematically attributable to electricity alone. To become carbon neutral by 2050 as outlined by the EU green deal there is still a long way to go. Solar is growing strongly in Germany, while onshore wind energy is happening too slowly though approvals have almost tripled along with grid expansions quadrupled. For the transition in energy demand sectors there are no structural reductions. Industry fluctuates with economic performance, buildings with the weather and subsidies for heat pumps, transportation also sees a reduction in electric car sales due to reduction of incentives. The German Emissions Trading Authority announced revenues from the sale of emission rights rose by 100 million euros to 18.5 billion euros in 2024. The revenues fill the climate and transformation funds, which subsidizes projects for the energy transition, such as building renovation, charging infrastructure for electric vehicles or hydrogen projects. The availability of wind and solar power is well balanced over a year’s period as wind is stronger during the winter and solar in the summer months. The volatility over a 1-2-week period is a challenge though and electricity storage is very important. So far batteries or hot water are used in buildings, but very few batteries in electric vehicles allow reversed charging. Hydro power storage is a well-established alternative, but an increase would mean large distortions of landscape. Underwater pumped storage power plants could provide a solution. They are based on water-filled hollow spheres sunk to the seabed, which are pumped empty using excess electricity from wind turbines, for example. The stored energy is used again by opening a valve and the inflowing water drives a turbine, which in turn generates electricity via a generator. Researchers from the Fraunhofer Institute for Energy Economics and Energy System Technology (IEE) in Kassel tested the technology with a three-meter concrete sphere that sunk 100 meters deep into Lake Constance. A concrete sphere with a diameter of nine meters is to be installed at a depth of 500 to 600 meters off the Californian coast near Los Angeles. In 2026, the first underwater pumped storage power plant in the Pacific with an output of 0.5 megawatts is to be operated on a trial basis for a year. Technology is ever evolving, the role of lubrication, friction and wear will change compared to the past, but remember Tribology is everywhere. Your editor in chief Manfred Jungk CO 2 -Emissions in Germany 3 % reduced Events 2 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 Events We look forward to your contribution! The scientific journal Tribologie und Schmierungstechnik (TuS) is one of the leading publications for tribological research in Germany, Austria and Switzerland. As the official journal of the Society for Tribology (GfT) in Germany, the Austrian Tribological Society (ÖTG) and Swiss Tribology, the issues provide information on research from industry and science, current events and developments in the specialist community. Further information on the journal and publication: https: / / elibrary.narr.digital/ xibrary/ start.xav? zeitschriftid=tus&lang=en Date Place Event ► 26.04. - 29.04.25 Copenhagen, Denmark ELGI 35 th Annual General Meeting ► 13.05. - 15.05.25 Brannenburg, Germany Oildoc Conference ► 18.05. - 22.05.25 Atlanta, Georgia (USA) 79 th STLE Annual Meeting & Exhibition ► 28.07. - 30.07.25 Zürich, Switzerland European Conference on Tribology - ECOTRIB ► 29.09. - 01.10.25 Wernigerode, Germany 66. GfT Conference Tribology Contents 3 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 Tribologie und Schmierungstechnik Tribology - Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie An Official Journal of Österreichische Tribologische Gesellschaft An Official Journal of Swiss Tribology Volume 71, Issue 5-6 March 2025 5 Timo Hacker, Arne Bischofberger, Katharina Bause, Sascha Ott, Albert Albers Optimization of the run-in and wear behavior of dry-running friction systems through the targeted adaptation of the counter-friction disk: Experimental investigations and potentials for sustainable and efficient solutions 14 Thomas Decker, Georg Jacobs, Julian Röder, Timm Jakobs, Jan Euler Derivation of running-in procedures for planetary journal bearings in wind turbine gearboxes by means of abrasive wear simulation 23 Johannes Wirkner, Mirjam Baese, Astrid Lebel, Charlotte Besser, Hermann Pflaum, Katharina Voelkel, Thomas Schneider, Karsten Stahl Impact of Water Contamination and Iron Particles on the Performance Loss of e-Drive Transmission Fluids in Wet Clutches 29 Thomas Decker, Georg Jacobs, Carsten Graeske, Pascal Bußkamp, Julian Röder, Tim Schröder Model uncertainty of a multiscale, elasto-hydrodynamic simulation method for the prediction of abrasive wear in journal bearings 38 Laura Stubbe, Yvo Stiemcke, Sarah Mross, Sarah Staub, Konrad Steiner, Kerstin Münnemann, Oliver Koch, Stefan Thielen A new rubber-lubricant compatibility test on a tribometer for radial shaft seals 45 Danka Katrakova-Krüger, Jonas Kotscha, Christoph Budach, Peter Erdmann Materials Investigation of Test Wings from LCPC Abrasiveness Test with Different Soils and Varying Test Duration 1 Editorial CO 2 -Emissions in Germany 3 % reduced 2 Events Science and Research 57 News Award of the Vogelphohl Medal of Honour to Prof. Dr.-Ing. Gerhard Poll Columns Preface For authors Authors of scientific contributions are requested to submit their manuscripts directly to the editor, Dr. Jungk (see inside back cover for formatting guidelines). Anzeige 4 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 BOOK RECOMMENDATION expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 \ 72070 Tübingen \ Germany \ Tel. +49 (0)7071 97 97 0 \ info@narr.de \ www.narr.de The conference provides an international exchange forum for the industry and the academia. Leading university researchers present their latest findings, and representatives of the industry inspire scientists to develop new solutions. Main Topics > Trends lubricants and additives > Automotive and transport industry > Industrial machine elements and wind turbine industry > Coatings, surfaces and underlying mechanisms > Test methodologies and measurement technologies > Digitalisation in tribology > Digital Tribological Services: i-TRIBOMAT > Sustainable lubrication Target Groups > Companies in the field of lubrication, additives and tribology > Research facilities Nicole Dörr, Carsten Gachot, Max Marian, Katharina Völkel 24th International Colloquium Tribology Industrial and Automotive Lubrication Conference Proceedings 2024 1st edition 2024, 279 p. €[D] 148,00 ISBN 978-3-381-11831-1 eISBN 978-3-381-11832-8 Introduction Nowadays, the demands placed on technical systems are constantly increasing. As a result, the optimization of dry-running clutches and brakes is important. In addition to performance, the focus is also on sustainability and resource efficiency. To exploit the full performance potential of these systems, as much of the intended friction surface as possible must be actively involved in the friction process. To this end, the surface profile of the friction partners is smoothed during the run-in process and a stable friction layer is formed [1-3]. These mechanisms optimize the transmission behavior and are crucial for improving the service life, performance and coefficient of friction stability of the friction systems [4-6]. Until now, research has mainly focused on improving the run-in behavior by modifying the friction lining [7]. This work sets a new accent by investigating the influence of the counter-friction disk on the run-in and wear behavior. The aim is to expand the understanding of the role of the counter-friction disk in friction contact and to exert a targeted influence on tribological behavior. This is achieved within the scope of this work through experimental investigations. Counter-friction disks made of the materials steel C45 and cast-iron material GGG40 are investigated. These were modified by the finishing processes nitrocarburizing and phosphatizing. Both mass-pressed organic friction linings and sintered metallic bronze-based friction linings are used as friction partners. The experimental results should enable a well-founded evaluation of the tribological properties of the counter-friction disks and show how this influences the run-in and wear behavior. The findings are relevant for shortening run-in times and reducing the associated costs for friction system manufacturers. As a result, a run-in system can be implemented directly at the customer. State of research Previous studies on the run-in behavior of dry-running friction pairings have mainly focused on the friction lin- Science and Research 5 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0033 Optimization of the run-in and wear behavior of dry-running friction systems through the targeted adaptation of the counter-friction disk: Experimental investigations and potentials for sustainable and efficient solutions Timo Hacker, Arne Bischofberger, Katharina Bause, Sascha Ott, Albert Albers* submitted: 17.09.2024 accepted: 21.11.2024 (peer review) Presented at GfT Conference 2024 This paper investigates the friction and wear behavior of dry-running friction systems during the run-in process, focusing on the specific adaptation of the counter-friction disk. The influence of these factors on the run-in behavior is investigated by varying the material and the finishing process of the counter-friction disk. The materials steel C45 and cast iron GGG40 as well as the application of nitrocarburizing and phosphating represent varied parameters. The results show that an improvement in wear properties, performance and coefficient of friction stability can be achieved by optimizing the counter-friction disk. These findings make it possible to increase the efficiency of friction systems. Keywords run-in Brake systems, System tribology, static coefficient of friction, dynamic coefficient of friction, wear coefficient, Phosphating counter-friction disk, Nitrocarburizing counter-friction disk Abstract * Timo Hacker, M. Sc. (corresponding author) Karlsruher Institut für Technologie, Institut für Produktentwicklung (IPEK), Kaiserstraße 10, Karlsruhe Arne Bischofberger, M. Sc.; Dipl.-Ing Katharina Bause; Dipl.-Ing Sascha Ott; Univ.-Prof. Dr.-Ing. Dr. h.c. Albert Albers; IPEK - Institut für Produktentwicklung am KIT - Karlsruher Institut für Technologie lationships between the coefficients of friction and the wear coefficient before and after run-in are investigated for different variants of counter-friction disks. A defined methodical procedure is developed for the experimental investigations with application-oriented load collectives. For this purpose, the following characteristic values are defined to quantify the influence of the counter friction disk: the static and dynamic friction coefficient and the wear coefficient. The tests are carried out at room temperature. The test sequence is divided into test series before the run-in and after the run-in to determine the static and dynamic coefficients of friction and the wear coefficient. At the beginning of the tests, the tribological system is examined in its new condition. For this purpose, measurements are taken to determine the wear and the static and dynamic coefficient of friction after the friction pairing has been installed. To keep the influence on the tribological system as low as possible, only three measurements are carried out to determine the static coefficient of friction in the new condition. During run-in, the friction pairings undergo a total of 1,000 braking cycles. The number of cycles for the run-in was determined based on the findings from preliminary tests. Once a certain number of braking cycles is reached under constant conditions, the coefficient of friction reaches an almost constant level [2]. After run-in, the measurements to determine the static and dynamic coefficient of friction and a final weight measurement are carried out again in the cooled system state Dynamic coefficient of friction Figure 1 shows the schematic test sequence of a braking cycle to determine the dynamic coefficient of friction. The drive accelerates the flywheel mass and the brake pad carrier to the specified speed and is then decoupled by opening the electromagnetic clutch. The test head is then closed via the axial-alternating slide, which leads to the rotating friction lining being braked to a standstill by the frictional contact with the counter-friction disk. Static coefficient of friction Figure 2 shows the schematic test sequence for determining the static coefficient of friction. First, the test head with friction lining is rotated load-free for a random period of time. This ensures that the same macroscopic surface properties do not meet and form the frictional contact during the next switching operation. An axial force is then applied, causing the friction lining to be pressed against the counter-friction disk until the nominal surface pressure is reached. The torque is then increased via the electric motor until the friction contact breaks loose. Once the test is complete, the axial force is reduced. Science and Research 6 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0033 ing [7]. The influence of different load collectives on the coefficient of friction and the surface change with different friction linings is analyzed. In accordance with [8], the run-in behavior is influenced by the load, the environment and also by the properties of the counter friction disk. However, the influence of the counter-friction disk on the tribological behavior of dry-running friction systems has not yet been sufficiently researched. According to Severin and Musiol, a friction layer forms during the run-in process, which protects the underlying layer from thermal overload [1].Gauger describes a quasi-stationary state of the friction layer that occurs during the run-in process, in which constant friction and wear properties prevail [9]. Trepte extends the evaluation of the run-in process by analysing wear particles, which show a stabilization of the particle size after several braking processes [2]. Investigations by Tsang, Jacko and Rhee show that brand-new pairings (organic friction lining with cast iron counter-friction disk) cause high wear during the formation of the friction layer, while the friction partners are smoothed [3].Neumann and Groganz as well as Bergheim show the influence of the load selective on the wear [5, 6]. In their work, Neumann and Groganz focus on the creation of a test collective with which the total wear in an application can be predicted as well as possible [5]. Bergheim concentrates on the influence of temperature as the central variable of the load spectrum and shows the influence of the geometry of the test specimens on wear [10]. Kleinlein is investigating the effect of different friction material combinations and environmental conditions on the friction and wear behavior of dry-running clutches by combining different friction linings with counter-friction disks in different materials. The counter-friction disk made of austenitic steel proves to be particularly wear-promoting. [11] Völkel investigated how the influence of steel plates affects the run-in behavior of wet-running clutches. The steel plates were produced using different manufacturing and finishing processes. The results show that the surface characteristics of the steel plates in particular have a significant influence on the run-in behavior of the friction pairing. [12] The previously cited studies show a correlation between the run-in and wear behavior of friction pairings through the design characteristics of the friction pairing. However, the previous studies did not focus on the influences of the counter-friction disks used for run-in. Research objective and method The experiments aim to gain initial insights into the influence of the counter-friction disk on the run-in and wear behavior of dry-running friction pairings. The re- Test environment The experimental investigations are carried out at the IPEK - Institute for Product Engineering at the KIT - Karlsruhe Institute of Technology. The test environment (Figure 3), which was developed according to the IPEK X-in-the-loop approach [14], is used to carry out reproducible tests on dry-running friction pairings under load collectives close to the application [15-17]. The inertias and stiffnesses of the residual systems from the application are mapped and the interaction effects with the residual system are represented realistically. To carry out the tests, the output side is mechanically blocked so that the tests are carried out as brake applications. During the tests, normal force, speed, torque, the travel distance of the axial force actuator and the temperature are measured. The applied axial force is recorded by a force sensor positioned between the test head and the axial force actuator. The transmitted torque is measured via a torque measuring hub, which is located as close as possible to the friction contact. This ensures that as little measurement distortion as possible is recorded due to other frictional torques, such as the bearings. The surface pressure, friction work, friction power, coefficient of friction and the gradient coefficient of friction are calculated from this data. To determine the temperature in the friction contact or the surface temperature, twelve thermocouples are inserted into the counter-friction disk. Results The results of the tests with nine different friction pairings are examined below. The different variants of the counter-friction disk are examined for their influence on the run-in and wear behavior. A mass-pressed organic friction lining is used as the friction partner for six pairings. For three pairings, it is investigated whether the influence of the counter-friction disk with a sintered metal bronze-based friction lining has a different behavior. The following friction pairings are investigated, see Table 1: Science and Research 7 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0033 Figure 1: Schematic test sequence for determining the dynamic coefficient of friction based on [7] Figure 2: Schematic test procedure for determining the static coefficient of friction based on [13] Static coefficient of friction Figure 5 shows the statistical evaluation of the static coefficient of friction. The box plot of the three repetitions of the static coefficient of friction measurement before the run-in is shown on the left and the ten measurements after the run-in are shown on the right. Most friction pairings show an increase in the static coefficient of friction after run-in. Friction pairings 2, 5 and 9 show a greater scattering of the measured coefficient of friction before run-in. The analysis of the static coefficient of friction of the friction pairings shows that the first coefficient of friction of each pairing exhibits the greatest deviation, while the subsequent values are close to each other. After the run-in, all friction pairings except friction pairings 8 show a small scatter. The outliers in friction pairings 1, 4 and 9 occurred at the beginning of the series of measurements. In the case of friction pairing 8, the large scatter is due to a drop in the static coefficient of friction towards the end of the test series. Science and Research 8 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0033 Dynamic coefficient of friction Figure 4 shows the statistical evaluation of the dynamic coefficient of friction. The box plot of the ten repetitions of the measurement of the dynamic coefficient of friction before the run-in is shown on the left and the ten measurements after the run-in are shown on the right. For friction pairings 1 and 8, the data shows very little scatter and a symmetrical distribution before the run-in. Friction pairings 2, 3, 4 and 5 show a moderate scatter around the mean value. However, the distribution is still relatively consistent. Friction pairings 6, 7 and 9 show a wider spread and possible asymmetry before the run-in. After the run-in, friction pairings 1, 2, 3, 4, 6 and 9 show an even distribution with few to no outliers. Friction pairings 5, 7 and 8 still show a moderate spread after the run-in. For most friction pairings, the mean dynamic coefficient of friction after run-in is lower than the values before run-in. This indicates an improvement in the active friction layer or a smoothing of the contact surfaces during run-in. Friction pairings 7 and 9 exhibit a higher mean dynamic coefficient of friction after run-in. Figure 3: Brake test bench for testing dry-running friction systems based on [7] Counter-friction disk friction lining friction pairing 1 C45 organic friction pairing 2 C45, nitrocarburized organic friction pairing 3 C45, phosphated organic friction pairing 4 GGG40 organic friction pairing 5 GGG40, nitrocarburized organic friction pairing 6 GGG40, phosphated organic friction pairing 7 C45, phosphated sintered metal friction pairing 8 GGG40, nitrocarburized sintered metal friction pairing 9 GGG40, phosphated sintered metal Table 1: Test pairings C45 organic C45, nitrocarburized organic C45, phosphated organic GGG40 organic GGG40, nitrocarburized organic GGG40, phosphated organic C45, phosphated sintered metal GGG40, nitrocarburized sintered metal GGG40, phosphated sintered metal organic organic organic organic organic organic sintered metal sintered metal sintered metal g g Science and Research 9 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0033 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 1 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 2 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 3 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 4 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 5 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 6 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 7 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 8 before run-in after run-in 0.2 0.25 0.3 0.35 0.4 0.45 0.5 dynamic coefficient of friction Friction pairing 9 Figure 4: Statistical evaluation of the dynamic coefficient of friction before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 1 before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 2 before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 3 before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 4 before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 5 before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 6 before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 7 before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 8 before run-in after run-in 0.15 0.2 0.25 0.3 0.35 0.4 0.45 0.5 static coefficient of friction Friction pairing 9 Figure 5: Statistical evaluation of the static coefficient of friction Most surfaces show an increase in roughness after exposure. An increase in the height amplitude is also observed, which indicates an increase in the peak depth and a change in the mean height Sv after stressing. For some surfaces, the differences in skewness (Ssk) show that the asymmetry of the surfaces has changed. The kurtosis also shows changes for most surfaces, with a tendency for the kurtosis to increase. Due to the finishing processes, both counter-friction disks show a smaller increase in the surface characteristics after run-in. Phosphating causes a smaller increase in the roughness of the C45 steel counter-friction disk than nitrocarburising. The reverse is true for the counterfriction disk made of GGG40 cast iron. When using the sintered metallic friction lining, the phosphated C45 counter-friction disk shows a similar development of the characteristic values as when using the organic friction lining. With the GGG40 counter-friction disk, the sintered metallic friction lining causes a reduction in the characteristic values. This indicates a stronger smoothing of the surfaces. Discussion Various findings can be derived from the experimentally determined dynamic and static coefficient of friction before and after the tun-in. By comparing the coefficient of friction at the different test times, it is possible to determine how the friction conditions have changed because of the run-in. The analysis shows that different pairings exhibit different wear behavior. Some pairings exhibit very high consistency and low wear rates, while others show significantly higher wear rates and greater variability. This may be due to the material composition, the surface treatment or the operating conditions of the friction surfaces. For most friction pairings, the mean dynamic coefficient of friction decreases after run-in. This indicates that the surfaces are better matched to each Science and Research 10 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0033 Wear coefficients Figure 6 shows the wear coefficients of the nine friction pairings across the run-in. Friction pairings 1 and 2 have the lowest wear coefficients and a low scatter. Friction pairs 2, 6, 8 and 9 show higher wear coefficients with moderate to low scatter. Friction pairings 3, 4 and 5 have higher wear coefficients with higher scatter. The finishing processes lead to an increase in the wear coefficient for the steel C45 counter-friction disk when using an organic mass-pressed friction lining (see friction pairings 1, 2 and 3). In the case of the phosphated steel C45 counter-friction disk, the sintered metal friction lining causes less wear than the organic friction lining (see friction pairings 2 and 7). In the case of the counter-friction disk made of cast iron material GGG40, nitrocarburizing (friction pairing 5) leads to an increase and phosphating (friction pairing 6) to a reduction in the wear coefficient when using a sintered metallic bronze-based friction lining. In the case of the counter-friction disk made of cast iron material GGG40, the finishing processes show a diametrical behavior of the wear when using the sintered metallic friction lining (friction pairing 8 and 9) compared to the organic friction lining (friction pairing 2 and 3). Surface analysis The figure shows the differences between the surface characteristics before and after the run-in of the nine friction pairings. A positive difference means that the value after the run-in is higher than before the run-in. Three measuring fields are defined on the counter-friction disc of the friction pairings. The position is programmed identically for each measurement. The measuring fields on the counter-friction disk have an area of 10 mm 2 and are offset by 120 degrees to each other. The measurements are carried out in new condition and after running-in. Friction pairing 1 Friction pairing 2 Friction pairing 3 Friction pairing 4 Friction pairing 5 Friction pairing 6 Friction pairing 7 Friction pairing 8 Friction pairing 9 0 50 100 150 Coefficient of wear in mm³/ MJ Figure 6: Statistical evaluation of the wear coefficient over the run-in other through the run-in process and a friction-active layer is formed, which leads to improved frictional efficiency. Friction pairings 7 (C45, phosphated) and 9 (GGG40, phosphated) show a higher mean dynamic coefficient of friction after run-in, which may indicate the formation of abrasive particles by the sintered metallic friction lining. In the case of the steel C45 counter-friction disk in combination with an organic friction lining, nitrocarburizing leads to a reduction in the dynamic coefficient of friction by around 5 % and the static coefficient of friction by around 27 %. This is accompanied by a significant increase in wear. In contrast, phosphating causes the dynamic coefficient of friction in the tests to remain at a similar level to the untreated counter-friction disk, while the static coefficient of friction increases by more than 15 %. However, the phosphating causes a significant increase in wear. In the case of the steel C45 counter-friction disk with a sintered metal friction lining, the dynamic coefficient of friction increases after run-in, while wear remains low. For the counter-friction disk made of cast iron material GGG40 with an organic friction lining, nitrocaburation leads to a lower dynamic coefficient of friction and increased wear in the tests. In addition, the static coefficient of friction is around 21 % lower and shows a large scatter before the run-in. Phosphating leads to a reduction in the dynamic coefficient of friction of around 5 % and the static coefficient of friction of around 4 %. However, it causes a reduction in wear. Nitrocarburizing shows an increase in the wear coefficient, while phosphating causes a reduction. This may indicate that the finishing processes influence the material properties differently. Phosphating appears to improve wear resistance, while nitrocarburizing may lead to increased wear. The finishing processes show different behavior when using sintered metallic and organic friction linings, indicating the importance of surface treatment and material selection. Science and Research 11 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0033 S q -2 0 2 4 6 8 μm S sk 0 5 10 S ku -100 0 100 200 S p -50 0 50 100 μm S v -100 0 100 μm S z -50 0 50 100 μm Measuring field 1 Measuring field 2 Measuring field 3 S a -2 0 2 μm Figure 7: Difference in the parameters of the surface analysis before and after the run-in S q S sk S ku S p S v S z S a S q S sk S ku S p S v S z S a [3] P. H. Tsang, M. G. Jacko und S. K. Rhee, “Comparison of chase and inertial brake dynamometer testing of automotive friction materials,” Wear, Jg. 103, S. 217-232, 1985. [4] H. Czichos und K.-H. Habig, Tribologie-Handbuch: Tribometrie, Tribomaterialien, Tribotechnik, 3. Aufl. Vieweg+Teubner Verlag; Springer Fachmedien Wiesbaden GmbH, 2010. [5] K. Neumann und O. Groganz, “Verschleißverhalten von organischen Kupplungsbelägen. Steigerung der Praxisnähe durch neue Prüfkollektive.,” (Kupplungen und Kupplungssysteme in Antrieben 2015), S. 257-268. [6] M. Bergheim, “Organisch gebundene Kupplungsbeläge Möglichkeiten und Grenzen,” VDI-Berichte, Jg. 1323, S. 527-548, 1997. [7] R. Fehrenbacher, A. Fischer, S. Ott, I. Cokdogru, D. Gierling und F. Mazurek, “FVA 860 I Abschlussbericht: Vorkonditionierung von trockenlaufenden Reibpaarungen für eine stabile Reibfunktion im Feld,” Nr. 1472, S.1-236, 2022. [8] L. Deters, A. Fischer, E. Santer und U. Stolz, “GfT Arbeitsblatt 7 - Tribolgoie: Verschleiß, Reibung - Definitionen, Begriffe, Prüfung,” S. 1-50, 2002. [9] D. Gauger, Wirkmechanismen und Belastungsgrenzen von Reibpaarungen trockenlaufender Kupplungen (VDI Reihe 1 Nr. 301) (Konstruktionstechnik/ Maschinenelemente). Düsseldorf: VDI Verlag, 1998. [10] P. J. Blau und B. C. Jolly, “Wear of truck brake lining materials using three different test methods,” Wear, Jg. 259, S. 1022-1030, 2005. [Online]. Verfügbar unter: http: / / dx.doi.org/ 10.1016/ j.wear.2004.12.022 [11] D. Severin und C. Kleinlein, “Einfluss der Reibwerkstoffkombination und der Umweltbedingungen auf das Reibverhalten trockenlaufender Kupplungen. Abschlussbericht zum Forschungsvorhaben 215/ II der Forschungsvereinigung Antriebstechnik e.V. (FVA), Heft 636,” 2001. [12] K. Völkel, H. Pflaum und K. Stahl, “Einflüsse der Stahllamelle auf das Einlaufverhalten von Lamellenkupplungen,” Forsch Ingenieurwes, Jg. 83, Nr. 2, S. 185-197, 2019, doi: 10.1007/ s10010-019-00303-2. [13] R. Fehrenbacher, S. Ott, I. Cokdogru und J. Huber, “FVA 806 I Abschlussbericht: Entwicklung einer Prüfmethode zur Ermittlung des statischen Reibmomentes (Losreißmoment) trockenlaufender Kupplungen und Bremsen auf Komponentenebene,” [14] A. Albers, M. Behrendt, S. Klingler und K. Matros, “Verifikation und Validierung im Produktentstehungsprozess,” in Lindemann (Hg.) 2016 - Handbuch Produktentwicklung, S. 541-569. [15] A. Albers, S. Ott und J. Kniel, FVA 607 II: Kupplungsmodell Trockenlauf II: Einfluss der Reibbelagsgeometrie auf das tribologische Verhalten: Forschungsvorhaben Nr. 607 II, 2015. [16] A. Albers, S. Ott und N. Schepanski, FVA 737: Leistungsgrenzen Trockenlauf: Methode zur Bestimmung der Leistungsgrenzen trockenlaufender Friktionssysteme, 2016. [17] T. Klotz, S. Ott und A. Albers, “Analyse des Schädigungs- und Erholungsverhaltens trockenlaufender Friktionspaarungen,” Forschung im Ingenieurwesen, Nr. 83, S. 209- 218, 2019. Science and Research 12 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0033 The results show that the properties of the counter-friction disks and the type of friction lining have a significant influence on the friction behavior and wear. Phosphating stabilizes the coefficient of frictions behavior in a large proportion of the friction pairs investigated and leads to a reduction in wear. Nitrocarburizing has different effects on wear depending on the friction lining. The material properties, the finishing processes and the use of different friction linings influence both the dynamic and static coefficients of friction as well as the wear rate. The knowledge gained enables a targeted, more efficient and sustainable design of future friction pairings, which can be optimally adapted to the respective applications. Outlook To further deepen the knowledge gained, the choice of material combination is to be expanded in future investigations. The combinations of steel C45 and cast iron material GGG40 with organic and sintered metal friction linings investigated so far provide valuable findings, but it remains to be seen how the tribological behavior of other potential material pairings will turn out under similar conditions. The influences of the finishing processes should also be investigated further to be able to control the run-in and wear behavior even more specifically. In addition, it makes sense to focus on the effects of the counter friction properties under different load collectives. Different load profiles, speeds and temperature scenarios could significantly change the friction and wear dynamics [5, 6, 9], which is of great importance in real application scenarios. Acknowledgement The investigations presented in the publication were performed within the FVA-Project 964 I. The authors acknowledge the funding of the research project. The FVA-Projekt 964 I “Einfluss Gegenreibscheibe” is funded by the “Forschungsvereinigung Antriebstechnik e.V. (FVA)“. Literature [1] D. Severin und F. Musiol, “Der Reibprozeß in trockenlaufenden mechanischen Bremsen und Kupplungen,” Konstruktion, Jg. 47, S. 59-68, 1995. [2] S. Trepte, “Einlaufverhalten technischer Reibbeläge,”Automobiltechnische Zeitschrift, Jg. 103, Nr. 12, 2001. Science and Research 13 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 More information and registration www.tae.de/ 50019 25 th International Colloquium Tribology Bridging Science and Industry - Driving a Sustainable Future with Tribology Join Europe`s leading conference on lubrication, friction and wear! Experience 3 intensive days featuring 130 presentations from top experts in research, industry and practice across 5 parallel sessions, attracting over 400 participants from around the globe. Don‘t miss the special 25th anniversary edition — save the date today! 27th - 29th January 2026 Ost昀ldern/ Stuttgart, Germany Submit your paper until 30th of April 25 planetary bearings in wind turbine (WT) gearboxes were published in 2015 [1]. Since then journal bearings play a crucial role in new generation wind turbine drivetrains. The transition from rolling element bearings to journal bearings enhances gearbox torque density due to the compactness of journal bearings. Furthermore, journal bearings promise a higher turbine reliability due to their potentially unlimited fatigue lifetime [2]. Especially in the light of great expenses for operation and maintenance (around 33 % of the levelized costs of electricity for offshore wind energy in 2021 [3]) the increase in drive train reliability remains a priority for the wind industry. The reliability of journal bearings is strongly influenced by the risk for a spontaneous failure. This can be prevented through a targeted running-in by abrasive wear. Wear is Science and Research 14 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 1 Introduction and Motivation In recent years, there has been a trend in wind energy technology to replace rolling-element bearings as planetary bearings in wind turbine gearboxes with journal bearings. The first groundbreaking investigations by M EYER et. al. on the suitability of journal bearings as Derivation of running-in procedures for planetary journal bearings in wind turbine gearboxes by means of abrasive wear simulation Thomas Decker, Georg Jacobs, Julian Röder, Timm Jakobs, Jan Euler* submitted: 20.09.2024 accepted: 28.12.2024 (peer review) Presented at GfT Conference 2024 * Thomas Decker, M.Sc. Prof. Dr.-Ing. Georg Jacobs Julian Röder, M.Sc. Timm Jakobs, M.Sc. Jan Euler, M.Sc. Chair for Wind Power Drives (RWTH Aachen University), Campus-Boulevard 61, 52074 Aachen The implementation of journal bearings as planetary bearings has enabled higher torque densities of planetary gear stages in wind turbines. Since journal bearings are generally more compact than rolling element bearings they enable more planetary gears to be fitted into a planetary gear stage compared to one of similar size with rolling element bearings. Additionally, when designed and operated correctly, journal bearings operate reliable with potentially unlimited fatigue lifetime. These advantages have led to a technology shift in recent years from rolling element bearings towards journal bearings as planetary bearings in wind turbine gearboxes. At the beginning of their service life journal bearings are subjected to a running-in phase either on a test rig or directly in the field. During the running-in phase the contour and surface roughness of the bearing adapt mainly due to abrasive wear in dependence of the loads the bearing experiences. The bearings must operate within their associated system environment (e.g. gearbox) so that the bearing properties can be established ideally. For optimal running-in of the bearings special load procedures can be executed in a controlled environment. Running-in procedures often in- Abstract clude gradually increasing loads until nominal load is achieved. For journal bearings in wind turbines this approach is not always applicable, since a deliberate running-in is a time-consuming process and therefore not necessarily part of standard end of line tests of wind turbine gearboxes. Wear simulation tools enable a derivation of ideal running-in procedures for example in terms of a reduced procedure duration and energy input. Such a simulation enables the assessment of changes in the microgeometry and surface roughness due to abrasive wear under the influence of operation under mixed friction conditions. With a methodically defined running-in process, the desired contour is created quickly without exceeding the bearings fatigue limits. Thus, this work presents a method for the derivation of running-in procedures for journal bearings based on abrasive wear simulations. Keywords journal bearings, wear simulation, wind turbines, planetary bearings, running-in the loss or degradation of material from the sliding surfaces of the journal bearings due to tribological stresses [4]. Among the known wear mechanisms [5], abrasion is best understood in terms of a simulative representation. With abrasive wear simulations the wear behavior of journal bearings can be investigated prior to expensive prototype testing. Therefore, wear simulations have the potential to play an important role in the industrial design process of journal bearings and the derivation of running-in procedures. S CHERGE et. al. demonstrated with an experimental study on tribometers that the long-term wear behavior of tribological systems can be improved by intentionally applying high stresses during the running-in process at the beginning of the bearing’s service life. Without a specific running-in process, the wear rate during the first hours of operation is lower than in the test with the running-in process. However, the wear rate of the specimen that previously has been run-in under high stress has proven to be lower. Thus, in the long run the bearing that was run-in shows lower wear [6]. In [7] an experimental study was presented showing the effect of a gradually decreasing sliding speed on the running-in behavior of journal bearings. It was demonstrated that a cyclic repetition of a certain running-in procedure leads to a reduction of measurable friction moment over time. Although the running-in of journal bearings has been well researched experimentally, the simulative evaluation of running-in procedures with varying loads is rarely discussed in the literature. This work presents a method for the derivation of running-in procedures for journal bearings based on abrasive wear simulations. Different running-in procedures are simulated and compared in terms of the generated wear pattern, asperity contact and generated friction. The aim of this analysis is a method that will allow for the derivation of an optimum running-in procedure in the long term. To achieve this an existing simulation tool for the calculation of abrasive wear is extended by the calculation of time-varying loads and a thermal model. The method is showcased using the model of a component test rig. A qualitative validation is presented using experimental results. Lastly a transfer of the simulation method to a full-size WT gearbox model is presented. 2 State of the art The calculation of abrasive wear on hydrodynamic journal bearings operating under mixed friction conditions has been subject of research and development for several years and is addressed in numerous publications [8, 9]. An overview of the established method for the simulation of abrasive wear used in the investigation in this paper is shown in Figure 1. Typically wear simulations consist of an iterative loop between an elasto-hydrodynamic (EHD) simulation (1) in combination with a contact model (2) and a wear calculation model (3) [9, 10]. The oil film height h i,j and the asperity contact pressure p ai,j are calculated using a commercially available EHD/ MKS tool comparable to the work by König et. al. [9]. With every iteration step i of the wear simulation a time increment t acc,i of the wear process is calculated, which in turn depends on the calculated wear rate dh W ⁄dt in this increment [11]. The wear height h W,i,j is calculated in (3) and looped back into the EHD simulation, where it is considered as a change in the bearing’s contour c B,i,j . This results in a change in asperity contact pressure p ai,j and hydrodynamic pressure p h,i,j . Due to the high computational effort required for EHD simulations, common wear simulations are based on calculating only one or a few revolutions of the bearing under static load in the EHD simulation until a steady state in terms of pressure and oil film height h i,j is reached. This steady state is then assumed to be constant over the entire duration of Science and Research 15 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 Termination EHD simulation Asperity wear model 𝑡 𝑎𝑐𝑐 ≥ 𝑡 𝑚𝑎𝑥 Oil film height ℎ 𝑖,𝑗 Asperity contact pressure 𝑝 𝑎,𝑖,𝑗 Wear height ℎ 𝑊,𝑖,𝑗 Surface roughness 𝑅 𝑎,𝑘 Wear height ℎ 𝑊,𝑖,𝑗 Contour wear model Contact model Wear height Hydrodynamic pressure Start 1 2 3 4 Figure 1: Multi-scale simulation flowchart for abrasive wear of journal bearings based on [11] and [9] Science and Research 16 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 the iteration step i of the wear simulation. After a certain number of iterations n i , a steady state in terms of the bearing profile c B and pressure is achieved. The method presented in Figure 1 contains a termination criterion t max which leads to the finalization of the wear simulation once a predefined accumulated wear time (Eq. 1) is reached. Hereby, the wear occurring within a specified duration can be calculated. Eq. 1 K ÖNIG et. al. introduced the additional representation of asperity wear (4) at each increment representing the smoothing of the bearing’s surface. An experimental validation of this approach on a component test rig for journal bearings is presented in [9]. H AGEMANN et. al. demonstrated the importance of structural deformations (e.g. gear ovalization, bending of the planetary pin) in a WT gearbox on the wear behavior of planetary journal bearings. To capture those deformation effects it is crucial to represent structural deformation by means of a finite element model or multi-body simulation (MBS) models coupled with the EHD representation of the sliding contact [10]. In [11] it was demonstrated by Decker et. al. that the wear simulation method shown in Figure 1 can correctly represent the effect of running-in of journal bearings. The simulation yields decreasing asperity contact as the wear progresses over time and converges towards zero when a moderate load is applied. This selfhealing effect forms the basis of running-in simulation. A useful application for the aforementioned type of abrasive wear simulation was showcased by JAITNER et. al. They simulated the occurring wear of planetary journal bearings in a WT gearbox during a targeted running-in process using an EHD/ MBS. For the running-in procedure they chose a step wise increase in input torque and demonstrated, that the wear volume V W converges towards a steady state [8]. L INJAMAA et. al. suggest running-in of the bearing using a step wise decrease in sliding speed while keeping the specific pressure constant [7]. Both approaches target running-in but with a different concept to increase the bearings load. In reality, there are load limits for journal bearings that must not be exceeded in order to avoid damage (spontaneous failure) [12]. Therefore, the running-in process cannot be carried out under arbitrarily high loads, but must be conservative and gradual. There is therefore potential for optimization between the best possible running-in in terms of the wear generated, the process duration and the friction energy applied. In summary, simulative methods for calculating the abrasive wear on journal bearings are well established and experimentally tested running-in procedures can be found in literature as well. Currently, there is no standardized method for the derivation of running-in strategies for journal bearings without the need for experiments. This paper presents such a method based on the abrasive wear simulation tool chain. The running-in ef- = , fect is evaluated in terms of achieved wear amount, remaining asperity contact and friction energy during the procedure. The following chapter will elaborate on the necessary implementations to the wear simulation for the purpose of this work. 3 Model and Method The method used in this work is based on the works by K ÖNIG et. al. [9] and D ECKER et. al. [11]. It was developed for the calculation of abrasive wear (contour and asperity/ roughness wear) on planetary journal bearings for wind turbines and qualitatively validated by means of experiments [11]. The wear rate: Eq. 2 is calculated according to the wear law by F LEISCHER [13] as a function of the local asperity contact pressure p a,i,j , friction coefficient μ a,i,j [14], relative sliding speed of the journal bearing v S and the so-called friction energy density e R . e R is the model and material specific wear coefficient introduced by F LEISCHER [9]. The EHD/ MBS model used in this work represents a small journal bearing test rig. The test rig and the simulation model are shown in Figure 2. The MBS/ EHD model consists of two flexible bodies (housing with the journal and the shaft). The model parameters are listed in Table 1. The wear simulation method [11] is adapted to feature the simulation of time-variant operating conditions enabling the simulative investigation of different runningin procedures with changing load conditions over time. This is done by introducing varying input load values (specific pressure p̅ i and sliding speed v S,i ) to the EHD/ MBS model at each iteration step i of the wear simulation. Furthermore, a thermal modelling approach is implemented to the wear simulation. The methods presented in [11] and [9] have in common that the bearing temperature Θ B is assumed to be constant over time. In reality, the bearing temperature can vary over time as a result of fluid and asperity contact friction. As the viscosity of the lubricant is temperature-dependent, the temperature has a considerable effect on the bearing’s operating behavior. Therefore, in [15] a numerical modelling approach for calculating the temperature distribution Θ i,j of a planetary journal bearing is suggested based on a heat balance of the oil film. Based on the suggestions in [15] a simple thermal model is introduced for this work which determines a global mass temperature of the journal bearing Θ B . Heat transfer into the bearing due to friction and a dissipation term modelling heat flow from the bearing to the surrounding are considered. The heat balance implemented to the wear simulation is based on Eq. 3: , , = , , , , Parameter Value Bearing width 30 Bearing diameter 120 Radial clearance 70 Bearing material 12 2 ( = 100 ) Planet gear material 42 4 ( = 210 ) Lubricant viscosity class ISO VG 320 (PAO) Eq. 3 with α B being the heat transfer coefficient between the bearing and its surrounding structure and m B · c p being the thermal capacity of the bearing. The friction energy entry into the bearing over the area of the sliding surface W R is modelled according to [15]: Eq. 4 The time integration of Eq. 3 yields the temperature curve Θ B (t) over time. Compared to other thermal models, this approach represents a major simplification, but offers the advantage of a low computational effort. 4 Results The simulation results of the different running-in procedures are discussed below. The procedures differ in the load profile (pressure and sliding speed). In order to achieve comparability between the procedures, the entire parameterization of the model is kept constant except for the procedure itself. All procedures have the same = = ( ) = duration (t = 10 h) and at the end one additional operating point (p̅ ̅ = 8 MPa, v S = 0.2 m/ s) is simulated as a reference in order to compare the values for the remaining contact pressure p a . Pressure variation: Two exemplary simulation results for running-in procedures are shown in Figure 3. First a step-wise increase in load p̅ ̅ at constant sliding speed v S (Figure 3 (a)) as suggested by J AITNER et. al. [8] and second a load ramp at constant sliding speed respectively (Figure 3 (b)). The procedures are comparable in terms of the time integral of the specific pressure p̅ int and sliding distance d according to Eq. 5 and Eq. 6. These two values are used as an equivalence criterion in this work. Eq. 5 Eq. 6 The results indicate an almost identical amount of wear in both procedures with roughly 0.73 mm 3 of wear volume V W . The asperity contact pressure p a shown in Figure 3 was averaged over all nodes of the EHD mesh of = ( ) = ( ) Science and Research 17 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 Figure 2: (a) Journal bearing component test rig, (b) MBS/ EHD model of the test rig Table 1: Model parameters of the MBS/ EHD journal bearing test rig model (a) (b) pattern can be achieved in a shorter time with a steeper ramp-shaped load increase than with a step function without exceeding a certain asperity contact or friction energy input. Speed variation: As an alternative to the variation in pressure (gradual load increase) shown above, runningin can be achieved by gradually reducing the sliding speed v S . In [7] L INJAMAA et. al. demonstrated experimentally that a step-wise reduction in sliding speed over several hours of operation leads to a reduction in friction. Science and Research 18 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 the sliding surface. For both simulations an equal initial temperature of 60 °C is assumed. The running-in procedure with the step-wise load increase results in a higher bearing temperature Θ B in the first 2.5 h of operation than the simulation with the load steps. In Figure 3 (a) it can be seen that each new load step results in a peak in asperity contact pressure p a and an increase in bearing temperature Θ B slightly delayed. This cannot be observed in the ramp-wise load increase, where the asperity contact and temperature curve are smoother. From this, the hypothesis can be derived that a comparable wear Figure 3: Results of the wear simulation for running-in procedures with a step-wise load increase (a) and a ramp wise load increase (b) (a) (b) Figure 4: Results of the wear simulation for running-in procedures with a step-wise reduction of the sliding speed (a) and a ramp wise sliding speed decrease (b) (a) (b) Similar to the procedures tested on journal bearings in [7] a wear simulation at constant specific pressure (p̅ = 8 MPa) is performed with a step-wise decrease in sliding speed from 0.2 m/ s to 0.1 m/ s over a duration of 10 h. The integral specific pressure p̅ int , sliding distance d and the reference operating point after 10 h of operation is chosen to be equal to the procedures shown above. Similar to Figure 3 a step-wise (see Figure 4 (a)) and ramp-wise decrease in sliding speed is simulated to compare both procedures (see Figure 4 (b)). Both simulations resulted in comparable amounts of wear (0.80 - 0.82 mm 3 ) It is noteworthy that the two procedures from Figure 4 (speed) generate about 10 % more wear than the procedures with pressure variation shown in Figure 3. Just like the load steps in Figure 3, the gradual reduction in sliding speed causes a peak in contact pressure at the beginning of each new operating point. A summary of the simulation results shown above is provided by Table 2. All procedures can be characterized by identical values for the sliding distance covered during the procedure d = 8640 m and the integral specific pressure p̅ int = 156 MPa · h. The running-in procedures are compared in terms of the remaining maximum asperity contact pressure after 10 h of running-in, the time integral of the friction power (cf. Eq. 4) and the generated wear. In general, all running-in procedures yield a significant reduction in asperity contact compared to the beginning of the procedure (cf. Figure 3 and Figure 4). Additionally, no striking difference can be observed between the results of the step and ramp simulations. Here the equivalence criterion from Eq. 5 and Eq. 6 appears to be applicable (asperity contact and generated wear are almost equal). In terms of generated friction energy, the ramp procedures are favorable (~10 % less friction energy). The speed variation (procedures 3 and 4) yields about 10 % higher amounts of wear which corresponds to the higher friction energy generated during the running-in procedure. The most significant difference lies in the remaining asperity contact which is smaller in the load variation simulation results (procedures 1 and 2). Procedures 1 and 2 yield smaller amounts of wear and a slightly smaller remaining asperity contact. In the light of these evaluation metrics the load varying with a load ramp function approach appears to be favorable over the speed varying approach. a Experimental validation In this work, as described above, the wear simulation method was extended by an analytical, thermal model and the option to calculate the abrasive wear under the influence of varying loads during running-in procedures. In this chapter a qualitative validation of these features is presented. For the sake of simplicity, the validation is only demonstrated for a running-in procedure with stepwise increasing load exemplarily. For the validation an experiment on a journal bearing component test rig was performed. A step-wise increasing load procedure is chosen exemplary and for the experiment the identical test setup was used as presented in [11]. In Figure 5 (a) the measurement results for the friction moment M Fr (t) and bearing temperature Θ B (t) over time are compared with the simulation results. In total 6 consecutive load steps were tested for a test duration of 2.1 h. During the experiment the sliding speed v S was kept constant at 0.2 m/ s. Science and Research 19 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 Procedure Variable Value 1 Load step function = . Contact pressure remaining after = : , ( = ) = 0.23 MPa , ( = ) = 0.79 MPa Integrated contact friction work: = 39.93 J Wear volume reached: = 0.72 mm 2 Load ramp function = . Contact pressure remaining after = : , ( = ) = 0.24 MPa , ( = ) = 0.80 MPa Integrated contact friction work: = 36.35 J Wear volume reached: = 0.72 mm 3 Speed step function = . Contact pressure remaining after = : , ( = ) = 0.36 MPa , ( = ) = 1.01 MPa Integrated contact friction work: = 56.09 J Wear volume reached: = 0.80 mm 4 Speed ramp function = . Contact pressure remaining after = : , ( = ) = 0.37 MPa , ( = ) = 0.98 MPa Integrated contact friction work: = 45.43 J Wear volume reached: = 0.82 mm Table 2: Summary of the simulation result for four different running-in procedures of the bearing’s axial contour c B before and after the experiment are shown in comparison to the contour generated from the wear simulation. Here the axial contour of the bearing is evaluated directly in the load zone. It can be seen that wear occurred during the experiment, since a significant change in the contour is measured at the bearing edges (x B = 0 mm and x B = 30 mm), which corresponds to the simulation result as well. The running-in procedure caused a recession of the contour of approximately 11 μm at x B = 0 mm. From the validation experiment it can be concluded, that the wear simulation of consecutively increasing load steps yields realistic and reasonable results. Although the approach for modeling the bearing temperature is comparatively simple and therefore more computational efficient compared to other state-of-the-art modeling schemes, it can qualitatively model the increase of the global bearing temperature. b Simulation of running-in procedures on system level The approach presented above was demonstrated on a small component test rig for journal bearings. The simulations showed that the gradual increase in load yields slightly more favorable results than the approach with decreasing sliding speed. The operating behavior of a journal bearing always depends a lot on its system environment. In the case of a planetary journal bearing the Science and Research 20 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 One important observation is, that in terms of the mean friction moment simulation and measurement are in good agreement. The typical running-in of the bearing at each load step can be recognized by the frictional moment M Fr converging from an initially high peak value to a stationary value. During the first four load steps this effect is significantly more pronounced in the simulation than in the experiment. One possible reason is a deviation in the transition point from pure hydrodynamic operation to mixed friction between the simulation and the experiment. Additionally, the measured friction moment shows higher peaks (around 10 Nm more at the last load step) than in the simulation and the measured friction decreases significantly faster than the simulation result. One possible explanation for this deviation is the constant wear coefficient used in Eq. 2. A possible remedy to correct this deviation is the use of an alternative wear law such as the wear law presented in [16] featuring a time dependent wear coefficient. In terms of the bearing temperature Θ B a qualitative agreement between the simulation and the measurement can be observed. The temperature measurement was performed with a temperature sensor 3 mm under the sliding surface in the load zone of the bearing. Both temperature curves from simulation and measurement increase from initially 20 °C to around 35 °C towards the end of the experiment. Slight deviations of the simulated temperature from the measurement can be observed (~ 5°). In Figure 5 (b) the measurements Figure 5: Validation results for a running-in procedure tested on a journal bearing component test: (a) time series of measured and simulated friction moment and bearing temperature, (b) comparison between the measured and the simulated bearing contour after running-in (a) (b) system behavior of the gearbox in terms of deformations plays a crucial role in the wear assessment (e.g. gear mesh influences, titling of the gear wheel and bending of the planetary pin). To showcase the simulation approach of running-in procedures from this work on WT gearbox level additional simulations are performed using an MBS/ EHD model of a planetary gear stage of a 850 kW WT gearbox from a Vestas V52 turbine. The model comprises flexible bodies of the housing, planet carrier, gears and the sun shaft (cf. Figure 6) [11]. The most important model parameters are listed in Figure 6. All the other parameters are kept equal to those described in Table 1. The MBS/ EHD model of the gearbox is coupled to the wear simulation according to Figure 1. Insights into the wear simulation with the gearbox model can be found in [11]. Two exemplary simulations are performed with an either step-wise (Figure 7 a) and ramp-wise (Figure 7 (b)) increase in input torque leading to an increasing pressure. Both procedures are equivalent in terms of p̅ int and d. The simulation results shown in Figure 7 indicate a comparable amount of friction in both procedures with approximately 9 mm 3 of wear volume V W . Contrary to the results from the journal bearing test rig shown above the initial asperity contact is much higher and the increase in temperature at the beginning of the running-in procedure is more pronounced. This can be explained by the strong edge loading of the journal bearing due to the tilting of the gear wheel. This results in higher initial asperity contact pressure and higher amounts of wear compared to the test rig model. Nonetheless both running-in procedures result in a significant reduction in asperity contact compared to the initial condition of the simulation. At only 4 %, the difference between the two procedures in terms of the friction energy generated in the gearbox model is lower than in the component model. This confirms the findings described in [10] that the journal bearing behavior is strongly system-dependent. 5 Discussion and Outlook Journal bearings are a driver of torque density and reliability of wind turbines. Therefore, they contribute significantly to the improvements in cost effectiveness of wind energy achieved in recent years. The performance of a journal bearing is highly influenced by the runningin behavior at the beginning of its service life. The intentional induction of abrasive wear during the running-in phase at the beginning of the bearing service life can be beneficial to the performance of the bearing. This work Science and Research 21 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 Figure 7: Exemplary results of the wear simulation for running-in procedures with a step-wise load increase (a) and a ramp wise load increase (b) Planet gear Planet carrier Gearbox housing Sun shaft Journal Bearing 𝐵 = 190 𝑚𝑚 𝐷 = ∅178 𝑚𝑚 Figure 6: MBS/ EHD model of a planetary gear stage of a WT gearbox (a) (b) [3] Stehly, T. u. Duffy, P.: 2021 Cost of Wind Energy Review. 2022 [4] Deters, L., Fischer, A., Santner, E. u. Stolz, U.: GfT Arbeitsblatt 7 - Tribologie. Verschleiß, Reibung (Definitionen, Begriffe, Prüfung). 2002 [5] Sommer, K., Heinz, R. u. Schöfer, J.: Verschleiß metallischer Werkstoffe. Wiesbaden: Springer Fachmedien Wiesbaden 2018 [6] Scherge, M., Shakhvorostov, D. u. Pöhlmann, K.: Fundamental wear mechanism of metals. Wear 255 (2003) 1-6, S. 395-400 [7] Linjamaa, A., Lehtovaara, A., Kallio, M. u. Léger, A.: Running-in effects on friction of journal bearings under slow sliding speeds, Bd. 234. 2020 [8] Jaitner, D., Schmelzle, B. u. Fiereder, R.: Assessment of the Wear Behavior of Journal Bearings within a Planetary Gear Stage of a Wind Turbine Transmission. In: Bearing World Journal Vol. 7. Frankfurt am Main: VDMA Verlag 2022, S. 7-12 [9] König, F., Ouald Chaib, A., Jacobs, G. u. Sous, C.: A multiscale-approach for wear prediction in journal bearing systems - from wearing-in towards steady-state wear. Wear 426-427 (2019), S. 1203-1211 [10] Hagemann, T., Ding, H., Radtke, E. u. Schwarze, H.: Operating Behavior of Sliding Planet Gear Bearings for Wind Turbine Gearbox Applications—Part II: Impact of Structure Deformation. Lubricants 9 (2021) 10, S. 98 [11] Decker, T., Jacobs, G., Graeske, C., Röder, J., Lucassen, M. u. Lehmann, B.: Multiscale-simulation method for the wear behaviour of planetary journal bearings in wind turbine gearboxes. Journal of Physics: Conference Series 2767 (2024) 5, S. 52012 [12] Forschungsvereinigung Antriebstechnik e.V.: FVA755 - Gleitlagerverschleißgrenzen II. Projektabschlussbericht. 2022 [13] Fleischer, G., Größer, H. u. Thum, H.: Verschleiß und Zuverlässigkeit. Berlin: VEB Verlag Technik 1980 [14] Offner, G. u. Knaus, O.: A Generic Friction Model for Radial Slider Bearing Simulation Considering Elastic and Plastic Deformation. Lubricants 3 (2015) 3, S. 522-538 [15] Prölß, M.: Berechnung langsam laufender und hoch belasteter Gleitlager in Planetengetrieben unter Mischreibung, Verschleiß und Deformationen. Dissertation. 2020 [16] Lijesh, K. P. u. Khonsari, M. M.: On the Modeling of Adhesive Wear with Consideration of Loading Sequence. Tribology Letters 66 (2018) 3 Science and Research 22 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0034 presents a method for the derivation of running-in procedures for journal bearings based on abrasive wear simulation. Different load scenarios are simulated and compared in terms of friction during the running-in procedure, generated wear and the remaining contact pressure after running-in. A reference operating point was selected to compare the asperity contact pressures after the procedures. The simulated procedures are equal in terms of integral specific pressure and sliding distance. The results indicate that the approach using a gradual increase in load is favorable over the approach with decreasing sliding speed, since it generated less wear while resulting in a lower asperity contact. In terms of generated friction energy the ramp-wise increase in load is favorable over the step-wise load increase. A qualitative validation of the simulation model in terms of friction moment, bearing temperature and the generated wear contour was experimentally achieved on a journal bearing test rig. The running-in procedures with load variation were applied on a WT gearbox model and proved a transferability of the findings from the component level simulation to the system level. Overall, it can be concluded on the basis of the available results that the load variation generates a slightly better wear pattern than the speed variation and is also preferable due to the friction energy generated. A ramp-shaped progression of the load or sliding speed is to be preferred in any case, as peaks in the asperity contact are caused by the steps. Increasing the slope of the ramps offers the potential to save test time in the procedures with the same wear result. Potential for optimization can be derived from this. Further optimization of the runningin procedures will be presented in a future paper. In addition, the use of a simplified temperature modelling approach is particularly advantageous with regard to the high computational effort associated with the wear simulations of the running-in procedures. Future work will address how thermal models with a higher fidelity can be made applicable to the presented method. Literature [1] Meyer, T.: Validation of journal bearings for use in wind turbine gearboxes. inFOCUS: WINDPOWER (2015) [2] Thys, T. u. Smet, W.: Selective assembly of planetary gear stages to improve load sharing, Bd. 87. 2023 Introduction The friction behavior of wet clutches is notably impacted by how the friction interfaces interact with the lubricant. A degradation or performance loss of the latter can cause adverse NVH effects, perceptible as shudder noise to the driver. In a worst-case scenario, this can damage the entire powertrain unit. The decisive factor for this adversity is the course of the coefficient of friction (CoF) versus the sliding speed. When a negative slope or a static CoF considerably surpassing the dynamic CoF is present, the control behavior deteriorates, leading to an increased tendency to shudder [3]. This behavior can be obtained from Figure 1. To avoid these adverse effects, oil additives, such as friction modifiers (FM), serve to shape the friction behavior of wet clutches [4-6]. However, in addition to the gradual aging of the lubricants due to continuous energy input over their lifetime, external factors, such as water contamination, can, due to physical adsorption between water molecules and additive elements, accelerate this performance loss [7-9]. Numerous studies have indicated that water impacts the friction behavior of wet clutches, leading to a higher static CoF and a negative gradient of the CoF versus the slip speed [7,10 -14]. Besides that, Yokomizo et al. [15] detected an increase in the static CoF due to iron contamination by comparing the friction behavior at the be- Science and Research 23 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0035 Impact of Water Contamination and Iron Particles on the Performance Loss of e-Drive Transmission Fluids in Wet Clutches Johannes Wirkner, Mirjam Baese, Astrid Lebel, Charlotte Besser, Hermann Pflaum, Katharina Voelkel, Thomas Schneider, Karsten Stahl* submitted 20.12.2023 accepted: 20.12.2024 (peer review) Presented at GfT Conference 2023 The present study delves into the connection between the driving comfort and safety of electric vehicles and the e-drive transmission fluid (ETF), examining how water influx arising from environmental factors and iron particles resulting from wear of various transmission components can lead to performance degradation in the lubricant. The performance loss is evaluated based on the changes in the friction behavior of wet multi-plate clutches. Previous studies on automatic transmission fluids (ATF) have shown that water and iron contamination contribute to the earlier failure of transmission components and tend to lead to NVH behavior (noise, vibration, harshness), particularly in wet clutch systems [1,2]. Hence, it is essential to expand this knowledge on e-drive lubricants to prevent premature damage to various transmission components by exchanging the lubricant. This paper provides insights into component test rigs and associated test methods to determine the performance loss of the e-drive transmission fluid in an application-related test environment. Exemplary results are shown from investigations with wet clutch parts from a transfer case used in serial automotive applications, considering the influence of water contamination and iron particles as well as their mutual interactions. Keywords NVH, Shudder, Water Contamination, Iron Particles, Lubrication, Additives, ICP-OES, Friction Behavior, Wet Clutch, ETF, E-axle Abstract * Johannes Wirkner, M.Sc. 1 (corresponding author) Dr.-Ing. Mirjam Baese 2 Dipl.-Ing. (FH) Astrid Lebel, M.Sc. 3 Mag. Dr. Charlotte Besser 3 Dr.-Ing. Hermann Pflaum 1 Dr.-Ing. Katharina Voelkel 1 Thomas Schneider, M.Sc. 1 Prof. Dr.-Ing. Karsten Stahl 1 1 Technical University of Munich (TUM), School of Engineering & Design, Department of Mechanical Engineering, Gear Research Center (FZG), Boltzmannstrasse 15, D-85748 Garching, Germany 2 Magna Powertrain GmbH & Co. KG, Industriestrasse 35, A-8502 Lannach, Austria 3 AC2T research GmbH, Viktor-Kaplan-Strasse 2/ C, A-2700 Wiener Neustadt, Austria which was operated in slip mode. In contrast to other testing methods, component test rigs combine the advantages of the application reference of field tests with a simultaneous reduction in ongoing testing costs. Figure 2 shows an overview of the test rig setup with its modules. The clutch disks used in transfer cases from serial automotive production are mounted within the inspection chamber onto the respective carriers. While the inner carrier, which is connected to a specially designed shudder unit to reduce the torsional stiffness of the shaft, rotates, the outer carrier is fixed to the housing. The lubricant is injected in a radial direction from the inside via an oil nozzle. A force-controlled piston applies the pressure hydraulically. The creep drive accelerates the friction disks in the engaged clutch state to a defined rotational speed, which is determined by an incremental encoder at the end of the shaft. A complete clutch pack comprises six steel and five friction disks, leading to ten friction interfaces. Both disks are mainly defined by their mean diameter of 116 mm and the radial width of 18 mm of the friction surface, respectively. A detailed description of the test parts can be found in [2]. Oil samples (0.8 L per test) of the lubricant L-401 that were previously damaged in test rigand vehicle-based endurance runs in E-axles are evaluated by the friction behavior of wet clutches. General technical data of the lubricant are listed in Table 1. Untreated rainwater is injected to replicate genuine environmental conditions. The respective element content of the oil samples is determined by inductively coupled plasma optical emission spectrometry (ICP-OES) after microwave digestion Science and Research 24 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0035 ginning and the end of a durability test. Moreover, a correlation between water contamination and wear debris, specifically iron particles, has been discovered, eventually resulting in a reduction of the CoF [16,17]. The awareness of these effects obtained from conventional ATFs can now be extended to E-axles and ETFs within the scope of the present investigations. Due to the high tribological requirements, wet clutches are the component for evaluating failure criteria. Experimental Setup and Methodology The experimental investigations in the presented study were carried out on the KLP-260 component test rig, Figure 1: Course of the CoF versus the slid-ing speed based on Naunheimer [3] Figure 2: Schematic setup of the KLP-260 component test rig based on Meingassner [18] Lubricant Kinematic viscosity at 40 °C Kinematic viscosity at 100 °C Density at 15 °C L-401 21.7 mm²/ s 5.0 mm²/ s 0.853 g/ ml Table 1: Technical data of the tested lubricant with nitric acid (iCAP 7400 ICP- OES Duo, Thermo Fisher, Waltham, Massachusetts, USA) [19]. The experimental testing aims to determine a critical water amount that can be added to the lubricant without leading to adverse shudder behavior of the clutch. Therefore, an iterative test procedure is developed. After a run-in phase of 100 cycles, a pre-defined amount of water is added to the test rig. The lubricant is circulated for 15 minutes within the test rig to establish a homogeneous distribution of the added water. Then, a sequence with pressure p = 1.0 N mm -2 and rotational speed Δn = 100 rpm consisting of three steady and transient slip shifts in each configuration is run, respectively. The added water is entirely evaporated by an additional 90 cycles of steady slip, and if the system does not show any adverse NVH effects, an increased amount of water is added. This iterative process is repeatedly executed until irreversible shudder behavior occurs. The details of this test procedure can be obtained from Figure 3. To quantify the performance of the respective oil samples with their respective pre-testing background, the parameter m µ that represents the gradient of the CoF versus the sliding speed is used. Due to its initially described relation to the shudder tendency of wet clutches (Figure 1), this value enables the monitoring of the lubricant status at the addition of different water amounts. m µ is determined by calculating the average gradient of the CoF versus the sliding speed in the range of a rotational speed ∆n = 5 … 25 %. The value is again averaged for statistical validation over the three transient slip cycles. The calculated energy input corresponds to the work performed by various machine elements during previously conducted test rig and vehicle-based endurance runs. It should be noted that the efficiency of the overall application must be subtracted from this energy input to ultimately deduce the energy input into the lubricant. However, this is of secondary relevance for the investigations presented below. Results and Discussion Figure 4 shows the relationship between the gradient of the CoF m µ , which was evaluated at different levels of water contamination. The individually colored data points refer to a specific energy input brought into the lubricant by the previously performed endurance runs, ranging from dark blue for low energy input to dark red for high values. Generally, a gradient reduction can be obtained by further increasing the added water amount. The corresponding values of 100 % energy input and iron content were chosen according to the specifications of the manufacturer, where exceeding these limits would result in an oil change. Furthermore, a higher energy input leads to a decrease of m µ , consequently contributing to earlier performance loss of the lubricant. At energies exceeding 420 % and water contamination of more than 350 %, a negative m µ prevails, resulting in an audible shudder of the wet clutch system. Considering the oil sample with the highest energy input, this behavior emerges even at 0 % of water addition. To trace the arc from the energy input to the iron content, which results from the wear of various machine elements in the E-axle, Figure 5 shows the characteristic value m µ again depicted versus the water contamination. In contrast to Figure 4, the individual graphs are ordered and plotted according to the iron content ‘Fe’ determined by ICP-OES at the end of the test-rig or vehicle-based endu- Science and Research 25 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0035 Figure 3: Schematic test procedure to investigate the influences of water contamination on pre-deteriorated lubricant samples wards the metal surface of the steel disk through the forces of adsorption [20]. Given their substantial polarity, water molecules assume a position at the steel disk, thereby obstructing the dedicated additives from performing their intended task. Iron particles are subject to similar mechanisms in the lubricant and thus contribute to impeding additives from forming a tribo-layer between the respective clutch disks through catalytic effects [21]. This is underlined by the observed effects, as shown in Figure 4 and Figure 5. Since the course of the respective graphs does not seem to vary greatly regardless of the listing of energy input or iron content, a correlation of the lubricant-specific values (Energy and Fe - referring to the sample state prior to the tests in KLP-260) is shown in Figure 6. An approximately linear relation (R 2 = 0.7270) between energy input and iron content prevails, particularly in the range of low pre-damage of the lubricant. With increasing test duration, the scatter increases, which can be attributed to several reasons. Firstly, measurement uncertainties arise since the method of sample collection, and the period between sampling and ICP-OES analysis Science and Research 26 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0035 rance run before evaluation in the KLP-260. Essentially, a similar finding to the previous illustration emerges: an increased iron content exceeding 200 % in the lubricant leads to an earlier performance loss of the latter. Furthermore, a second finding can be derived from the two figures. The addition of water only seems to have an influence on the performance of the lubricant in the case of slight prior damage. In contrast, higher conditions of pre-damage (Energy > 300 %, Fe > 150 %) lead to an earlier lubricant deterioration and thus, shudder behavior of the wet clutch system even at low levels of water contamination. Therefore, it can be concluded that in an already pre-damaged state of the lubricant m µ does not further deteriorate through the addition of water. The changes in the friction behavior arise from the deterioration of the lubricant caused by the presence of water contamination in correlation with iron particles. Typically, the molecular makeup of additives such as friction modifiers, dispersants, or detergents comprises a polar head and a hydrocarbon tail. Under uncontaminated conditions, the polar head of friction modifiers is drawn to- Figure 4: Gradient of the CoF versus water contamination with respect to energy input Figure 5: Gradient of the CoF versus water contamination with respect to iron content significantly influence the measurement outcome as wear particles are deposited rapidly. However, the sampling within the scope of these investigations was kept constant allowing for improved comparability of the executed tests. Furthermore, differences e.g., in temperature levels of the test rig or vehicle-based endurance runs, have not yet been considered. Nevertheless, this regression enables abstracting and reducing the complexity of the mechanisms for future studies. Additionally, this visualization reveals a gap in the investigated oil samples between 200 % and 450 % energy input and highlights the potential for further investigations. Conclusion The present study focuses on the influences of energy input and iron particles with respect to water contamination of the lubricant by evaluating the friction behavior of wet clutches. Therefore, oil samples from previously performed test rig and vehicle-based endurance runs are tested in the KLP-260 component test rig. To quantify the performance loss of the lubricant, the characteristic value m µ , representing the gradient of the CoF versus the sliding speed, which serves as a performance indicator of the lubricant. The following conclusions can be drawn: • Due to physical adsorption and catalytic effects, water, and iron contamination significantly impair the performance of the lubricant. • Iron contents over 150 % and energy input exceeding 300 % lead to an earlier lubricant deterioration and, thus, shudder behavior of the wet clutch system. Above these values, additional water contamination does not further downgrade the performance of the lubricant. • A linear regression between energy input and iron content can be drawn to reduce the complexity of the systems approach. Acknowledgement This work was funded by the “Austrian COMET-Program” (project InTribology1, no. 872176) via the Austrian Research Promotion Agency (FFG) and the federal states of Niederösterreich and Vorarlberg. References [1] S. Stehle, M. Baese, K. Voelkel, K. Stahl, Determination of Friction and Shudder Behaviour of Slip-Controlled Wet Multi-Plate Clutches, Graz/ Spielberg, 2022. [2] J. Wirkner, M. Baese, A. Lebel, H. Pflaum, K. Voelkel, L. Pointner-Gabriel, C. Besser, T. Schneider, K. Stahl, Influence of Water Contamination, Iron Particles, and Energy Input on the NVH Behavior of Wet Clutches, Lubricants 11 (2023) 459. https: / / doi.org/ 10.3390/ lubricants11110459. [3] H. Naunheimer, B. Bertsche, J. Ryborz, Fahrzeuggetriebe: Grundlagen, Auswahl, Auslegung und Konstruktion, third. Auflage, 2019. [4] R. Maeki, B. Ganemi, E. Hoeglund, R. Olsson, Wet Clutch Transmission Fluid for AWD Differentials: Influence of Lubricant Additives on Friction characteristics, Lubr. Sci. 19 (2007) 87-99. https: / / doi.org/ 10.1002/ ls.33. [5] H. Spikes, The History and Mechanisms of ZDDP, Tribol Lett 17 (2004) 469-489. https: / / doi.org/ 10.1023/ B: TRIL.0000044495.26882.b5. [6] H. Ohtani, R.J. Hartley, D.W. Stinnett, Prediction of Anti- Shudder Properties of Automatic Transmission Fluids using a Modified SAE No. 2 Machine, SAE Ttransactions 103 (1994) 456-467. [7] N. Fatima, P. Marklund, R. Larsson, Water Contamination Effect in Wet Clutch System, Proceedings of the Institution of Mechanical Engineers, Part D: Journal of Automobile Engineering 227 (2013) 376-389. https: / / doi.org/ 10.1177/ 0954407012455145. [8] N. Fatima, Impact of Water Contamination and System Design on Wet Clutch Tribological Performance. Dissertation, Lulea, 2014. Science and Research 27 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0035 Figure 6: Linear regression of iron content and energy input [16] Y. Akita, K. Abe, Y. Osawa, Y. Goto, Y. Nagasawa, N. Sugiura, S. Wakamatsu, K. Kosaka, Analysis of Friction Coefficient Variation with Moisture between Friction Surfaces, SAE Technical Papers (2016). https: / / doi.org/ 10.4271/ 2016-01-0411. [17] A. Dorgham, A. Azam, A. Morina, A. Neville, On the Transient Decomposition and Reaction Kinetics of Zinc Dialkyldithiophosphate, ACS Applied Materials & Interfaces 10 (2018) 44803-44814. https: / / doi.org/ 10.1021/ acsami.8b08293. [18] G.J. Meingaßner, H. Pflaum, K. Stahl, Test-Rig Based Evaluation of Performance Data of Wet Disk Clutches, 14th International CTI Symposium (2015). [19] A. Agocs, A.L. Nagy, Z. Tabakov, J. Perger, J. Rohde- Brandenburger, M. Schandl, C. Besser, N. Dörr, Comprehensive Assessment of Oil Degradation Patterns in Petrol and Diesel Engines observed in a Field Test with Passenger Cars - Conventional Oil Analysis and Fuel Dilution, Tribology International 161 (2021) 107079. https: / / doi.org/ 10.1016/ j.triboint.2021.107079. [20] J. Crawford, A. Psaila, Miscellaneous Additives, in: R.M. Mortier, S.T. Orszulik (Eds.), Chemistry and Technology of Lubricants, first ed., Springer Science+Business Media, 1994, pp. 160 -173. [21] P. Nieuwland, T.A. Droste, Automatic Transmission Hydraulic System Cleanliness - The Effects of Operating Conditions, Measurement Techniques and High-Efficiency Filters, in: SAE Technical Paper Series, SAE International400 Commonwealth Drive, Warrendale, PA, United States, 2001. Science and Research 28 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0035 [9] N. Fatima, P. Marklund, A.P. Mathew, R. Larsson, Wet Clutch Friction Interfaces under Water-Contaminated Lubricant Conditions, Tribology Transactions 59 (2016) 441-450. https: / / doi.org/ 10.1080/ 10402004.2015.1081998. [10] H. Wang, Effects of Water in All-Wheel-Drive Clutch Oil on All-Wheel-Drive Cornering Noise, SAE Technical Papers (2022). https: / / doi.org/ 10.4271/ 2022-01-5103. [11] M. Baese, A. Lebel, R. Franz, J. Prost, Analysis of Wet Vehicle Clutch Damage Mechanism caused by Water, Tribologie und Schmierungstechnik 69 (2022) 7-13. https: / / doi.org/ 10.24053/ TuS-2022-0008. [12] W. Williamson, B. Rhodillegible, The Effects of Water on Cellulose-based Frictional Surfaces in Automatic Transmission Clutch Plates, SAE Technical Papers (1996). https: / / doi.org/ 10.4271/ 961917. [13] K. Berglund, P. Marklund, R. Larsson, M. Pach, R. Olsson, Wet Clutch Degradation Monitored by Lubricant Analysis, in: SAE Technical Paper Series, SAE International400 Commonwealth Drive, Warrendale, PA, United States, 2010. [14] A. Dorgham, A. Azam, P. Parsaeian, T. Khan, M. Sleiman, C. Wang, A. Morina, A. Neville, Understanding the Effect of Water on the Transient Decomposition of Zinc Dialkyldithiophosphate (ZDDP), Tribology International 157 (2021). https: / / doi.org/ 10.1016/ j.triboint.2021.106855. [15] M. Yokomizo, T. Iwai, K. Narita, M. Kudo, Lubricants Formulation Technology for Fuel Saving Performance in Automatic Transmissions, SAE International, 2015. 1 Introduction Wind energy production is a cornerstone in the transition from fossil energy to renewable energy. With a share of over 30 % in the year 2023 wind energy has already become the most important renewable energy source in Germany [1]. To further increase the energy production from wind and maintain its competitiveness wind turbines are designed with increased rotor diameters and thus higher power ratings [2]. Larger rotor diameters lead to an increase in weight of the rotor and to increased torque as well as non-torque loads that the drivetrain has to bear. Thus, the main bearing, the gearbox and the adjacent structure need to increase in size to carry the loads. The gearbox is one of the major contributors to the weight of a wind turbine drivetrain. The ring gear of the first planetary stage massively contributes to the gearbox weight and size. Thus, the size of the planetary stage needs to be decreased to increase the power density of the drivetrain. The use of journal bearings as planetary bearings instead of conventional rolling bearings is a recent driver of a decrease in weight of wind turbine gearboxes [3]. Journal bearings have a smaller installation space compared to rolling bearings since there are no rollers. In addition, journal bearings have an infinite lifetime and are remarkably reliable when designed and operated correctly. Special load cases, e.g. strong deformation due to overload, can eventually lead to mixed friction and thus abrasive wear in the journal bearings resulting in a change in contour and roughness of the sliding surface. For a re- Science and Research 29 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 Model uncertainty of a multiscale, elasto-hydrodynamic simulation method for the prediction of abrasive wear in journal bearings Thomas Decker, Georg Jacobs, Carsten Graeske, Pascal Bußkamp, Julian Röder, Tim Schröder* submitted: 20.09.2024 accepted: 11.01.2025 (peer review) Presented at GfT Conference 2024 Journal bearings have a potentially unlimited service life. In combination with the increased power density compared to rolling bearings journal bearings are attractive for the use in wind energy gearboxes. The reliability of journal bearings is significantly influenced by their wear behavior. Journal bearings are typically designed using elasto-hydrodynamic (EHD) simulations. The abrasive wear behavior of a journal bearing can be calculated with a wear algorithm which is coupled to the EHD-simulation. This approach is commonly used to simulate the abrasive material removal and smoothing of the surface roughness. State of the art wear calculation methods contain a large number of parameters and are prone to calculation errors due to parameter uncertainty. Therefore, the aim of this work is a quantified maximum model error (deviation of simulation results due to uncertainties in the parameterization) of a simulation tool for the calculation of abrasive wear in hydrodynamic journal bearings based on the wear model according to Fleischer. First, the influence of individual input parameters on the elasto-hydrodynamic simulation and the accumulated wear volume and wear depth is analyzed by means of a sensitivity analysis. The parameters with a relevant influence are identified and discussed on the basis of the sensitivity analysis. From the experimental work on a journal bearing test rig measurement uncertainties in terms of wear volume are derived and their influence on the wear coefficient and the overall model error are examined. The overall accuracy of the wear simulation model is then evaluated with regard to the selected input parameters in terms of a worst-case scenario analysis. Keywords Wear simulation, Journal bearings, Abrasion, EHDsimulation, Sensitivity analysis Abstract * Thomas Decker, M.Sc. Prof. Dr.-Ing. Georg Jacobs Carsten Graeske, M.Sc. Pascal Bußkamp, M.Sc. Julian Röder, M.Sc. Chair for Wind Power Drives (RWTH Aachen University), Campus-Boulevard 61, 52074 Aachen Dr.-Ing. Tim Schröder Vestas Nacelles Deutschland GmbH, Martin-Schmeißer-Weg 18, 44227 Dortmund 3 Method This work addresses the wear simulation method presented in [9] with the aim to quantify the maximum calculation uncertainty by means of a worst-case scenario analysis. The study is carried out using an EHD model of a journal bearing component test rig. The test specimens are specified in section 4.3 and more details about the test rig can be found in [9,11]. The calculation uncertainty is defined as the difference between a maximum and a minimum wear calculation result depending on two different sets of input parameters: one for a minimum and one for a maximum wear rate. Both extreme parameter sets consist of the most influential parameters at their lower and upper boundaries of a determined range based on literature. The most influential parameters in terms of wear behavior are identified using a sensitivity analysis. The sensitivity analysis is performed according to M ORRIS [12] . The wear simulation method was checked for plausibility in [9] by comparing it with experimentally determined wear depth values after tests on the journal bearing test rig. The applied wear simulation method can be divided into three parts (see also Figure 1). First, the EHD simulation model (1) calculates the asperity contact pressure p a,i,j for each node (i,j) considering the contact model according to G REENWOOD AND T RIPP [13] and taking additionally micro-hydrodynamic phenomena according to P ATIR AND C HENG [14] into account. The EHD model also considers general properties of the tribological contact (e.g. clearance S, lubricant viscosity) and operating conditions (specific pressure p̅ and sliding speed v S ). Secondly, based on the calculated asperity contact pressure distribution p a,i,j the wear model (2) calculates the wear depth per node h W,i,j according to Fleischer [10] considering a local coefficient of friction μ a,i,j according to O FFNER AND K NAUS [15]. Thirdly, the surface roughness model (3) calculates the surface smoothing according to D ECKER AND K ÖNIG [9,8]. Additionally, adjustments of the bearing surface profile lead to changes in the roughness. The updated roughness parameter R and the wear depth per node h W,i,j are used as input for the next iteration step until the simulation end time is reached. Figure 1 summarizes the wear simulation method as a flow chart, illustrates the models used in each section of the method and lists the parameters analysed within the sensitivity analysis. Table 1 gives an overview of the analyzed model parameters and the defined ranges based on literature, supplemented by measurements. For the roughness orientation Γ and elastic factor K recommended values from literature are used [17,16,7,19]. The deviation of the surface roughness R was determined by means of tactile surface measurements of the specimen before the experiment. This measurement study showed a variation in surface roughness of ± 10 % (see section 4.2). The deviation in the bearing temperature T stems from different friction intensities during the running-in process, which results in a decay of the temperature towards a steady state. More details about the ther- Science and Research 30 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 liable design of journal bearings simulations can be applied to assess the abrasive wear behaviour early in the design process without the need for expensive prototype testing. The major downside of state-of-the-art wear simulations is the high number of input parameters which makes the results prone to measurement and parameterization imprecisions. This work addresses the uncertainty of a given wear simulation method and suggests an approach to quantify the maximum calculation error. The first step is to identify the most influential parameters by means of a sensitivity analysis. The most important parameter is then determined metrologically using wear experiments on a journal bearing test rig and parameter inaccuracies due to measurement errors are also identified. The final result of this work is a worst-case quantification for the uncertainty in the wear simulation results. 2 State of the art for the simulation of abrasive wear Wear simulations have been subject to numerous research efforts in the past. P RÖLß [4] simulated the abrasive running-in wear behaviour of journal bearings in planetary gearboxes to analyse changes in contour and roughness. H AGEMANN ET. AL. [5] developed a wear calculation method for journal bearings of planetary gears considering elastic deformations and abrasive wear and D ING ET. AL. [6] presented a validation of their journal bearing simulation tool. K ÖNIG ET. AL. [7] demonstrated that abrasive wear on hydrodynamic journal bearings can be calculated reliably for constant and transient operating conditions (e.g. start-stop). Validation of this approach was achieved from experiments on a component test rig [8]. D ECKER ET. AL. [9] presented a multi-scale wear simulation method suitable for the simulation of abrasive wear in wind turbine journal bearings as an adaption of the work by K ÖNIG . The method consists of an iterative loop between an elasto-hydrodynamic (EHD) model of a journal bearing and a software tool simulating abrasive wear in terms of material removal and smoothing of the surface roughness. F LEISCHER’S wear law [10] is used for in this approach and supplemented by the smoothing of surface roughness. A validation of the method on a component test rig showed a good agreement between measurement and simulation [9]. Transferability of the wear simulation tool to different EHD models (e.g. planetary journal bearings in wind turbine application) was also demonstrated. L EHMANN ET AL. [11] evaluated different wear models for the wear calculation on journal bearings. All the aforementioned wear calculation approaches have in common that they have a high number of input parameters. This results in a high parameterization effort and the risk of calculation errors due to inaccuracies in the parameterization. There is a lack of studies investigating these inaccuracies. mo-hydrodynamic behavior of journal bearings can be found in [18]. Since the presented wear calculation method in this work assumes isothermal behavior of the bearing over time a constant bearing temperature is assumed. The sensitivity analysis examines the maximum and minimum measured temperature value during the experiments. Analogous, the bearing clearance S is kept constant and is assumed to have a range of uncertainty of ± 10 μm due to thermal expansion effects and deviations in the manufacturing process. The coefficient of friction is examined in a range typical for journal bearings [7]. The friction energy density e R of F LEISCHER’S wear law is examined in a range according to comparable experimental studies [17,16,19]. The parameter range for b, c and L S of the O FFNER AND K NAUS model is based on literature values [17,16]. The sensitivity analysis in this work is based on a factorial sampling plan according to M ORRIS [12] to rank the influence of the analyzed parameters on the outputs of the wear simulation method. Compared to other sensitivity analysis methods, M OR- RIS’ approach requires a comparably low number of simulations to analyze a high number (n = 10) of parameters [20]. The interpretation of the results is based on the mean μ *EE and the standard deviation σ EE of the so-called elementary effects (EE) of the parameters on the model output [12]. The quality of the results depends on the number of repetitions r which are required to calculate μ *EE and σ EE . S ALTELLI ET AL . [20] recommend r min = 4 as lower boundary to achieve sufficient results. M ORRIS ’ sampling method is based on one-factor-at-a-time randomization and restricted to analysis of elementary effects. Further details about designing the sampling matrices are given in [12]. Typically, the influence of the parameters on an output are interpreted graphically by plotting the standard deviation of the elementary effects σ EE over the mean elementary effects μ *EE . High mean values indicate a high influence of the corresponding parameter. Thus, analyzing the mean elementary effects enables a ranking of the parameter influence. High standard deviation values indicate a nonlinear effect [12]. 4 Results First, in section 4.1 the results of the sensitivity analysis are presented. The most influential parameters according to the resulting ranking are varied to calculate the worst- Science and Research 31 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 MBS/ EHD Simulation 1 Extended Reynolds equation to consider microhydrodynamics according to [Patir&Cheng] = Roughness orientation Contact model according to [Greenwood&Tripp] = . , K Elastic factor Roughness parameter General simulation parameter T Temperature Bearing clearance Coefficient of friction Wear model Wear model according to [Fleischer] 2 e R Friction energy density Local coefficient of friction according to [Offner&Knaus] , , = + , , , + , , , , , , with , , = , , b Model coefficient Model coefficient Reference length Surface model Surface smoothing per patch according to [König] 3 Asperity contact pressure , , Asperity contact ratio , , Wear depth , , Roughness Figure 1: Flow chart of the wear simulation method according to [9] Parameter Unit Lower boundary Upper boundary Reference Roughness orientation - 1 100 [17,16] Elastic factor - 0.0003 0.003 [17,7] Roughness parameter 10 % + 10 % measurements Temperature ° 55 70 measurements bearing clearance 0.14 0.16 measurements Coefficient of friction - 0.05 0.2 [17,7] Friction energy density 1 10 13 1 10 16 [17,16,19] Model coefficient (O&K) - 1,000 10,000 [17,16] Model coefficient (O&K) - 50 1,000 [17,16] Reference length 1 10 -6 2 10 -6 [17,16] Table 1: Parameter range for sensitivity analysis the typical representation of σ EE over μ *EE for each objective. The high μ *EE values of the friction energy density e R indicate the highest sensitivity of wear depth h W , wear volume V W , hydrodynamic pressure p h,e and asperity contact pressure p a,e at simulation end to the friction energy density e R . However, the friction energy density e R shows less significant influence on the hydrodynamic pressure p h,s and asperity contact pressure p a,s at the beginning of the simulation. The graphical determination of the parameters with the second and third largest influence on the wear behavior is not possible without further ado. To perform a rank-based evaluation, the parameters are ranked based on their μ *EE values for each objective. Subsequently, the mean rank regarding the objectives h W , V W , p h,e , p a,e is determined to evaluate the influence of the parameters on the wear and the pressure distributions at simulation end (Table 2). According to the determined mean ranks of the grouped objectives, the elastic factor K from the contact model by G REEN- WOOD AND T RIPP [13] is one of the two most influential parameters. The other second most influential parameter b and the following highest ranked parameters c and L S are coefficients of the local friction model according to O FFNER AND K NAUS [15]. The results show a significant influence of the parametrization of the contact model and the local friction model compared to the remaining parameters. Science and Research 32 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 case uncertainty of the wear simulation method. The friction energy density e R is the most significant parameter regarding wear. Due to high deviations (three orders of magnitude) of e R in the literature the parameter range is determined based on in-house experiments to increase the accuracy of the worst-case uncertainty calculation. The determination of e R is based on the measured wear in the experiments. Hence, in section 4.2 the wear volume measurement method and the determination of its uncertainty is presented. Section 4.3 explains the determination of e R for different test procedures. 4.1 Sensitivity analysis The Sampling for the sensitivity analysis is based on r = 5 repetitions to vary the parameters given in Table 1 on p = 4 levels. Thus, N = (n + 1) · r = 55 wear simulations are performed to analyze the sensitivity of the aforementioned parameters according to M ORRIS . The influence of the variation of the ten parameters on six objectives is evaluated. The considered objectives are the wear depth h W , wear volume V W , hydrodynamic pressure p h,s at simulation start, hydrodynamic pressure p h,e at simulation end, asperity contact pressure p a,s at simulation start and asperity contact pressure p a,e at simulation end. For each objective, the normalized mean μ *EE and the standard deviation σ EE of the elementary effects of the ten parameters on the model output is calculated. Figure 2 illustrates the results of the sensitivity analysis with Figure 2: Graphical results of the sensitivity analysis according to Morris for selected objectives Overall rank 1 2 2 4 5 6 6 8 9 10 Mean rank ( , , , , , ) 1.0 3.3 3.3 3.8 5.3 6.0 6.0 8.0 8.5 10.0 Parameter Table 2: Mean ranks of the parameters regarding wear and pressure distribution at simulation end As the friction energy density e R has the highest influence on the resulting wear volume V W and the wear depth h W , it is important to determine e R precisely. Since the values for the friction energy density e R are not known a priori, experimental investigations provide a good starting point for determining the relevant range of e R values for the operating points to be simulated. After the experimental determination of the parameter range for the conditions on the test bench, the influence of the uncertainty of the friction energy density e R on the wear behavior can be investigated. 4.2 Wear volume measurement uncertainty This work addresses the calculation uncertainty of the wear simulation method presented in [9]. However, the experimental measurement of the wear volume V W shows uncertainties too. The measured wear volume V W is used to determine the range of the friction energy density e R for the worst-case scenario analysis of the test rig wear simulations. Thus, the wear volume measurement uncertainty influences the determination of e R . The determination of the wear volume measurement uncertainty allows the precise identification of the possible range of e R . The wear volume V W is determined by a comparison of the contour before and after the experiment. The contour is determined using a tactile measurement in axial direction at seven angular measurement points ϕ i t [90°; 135°; 160°; 180°; 200°; 225°; 270°] of the journal bearing. Due to their availability and the low technical complexity tactile measurements are preferred in this work over other measurement approaches (e.g. optical 3d-contour measurements as suggested in [16]). In addition, in this work the friction energy density e R is determined directly on the basis of test bench tests, whereas in comparable work it was determined on tribometers [16]. The seven measurement positions ϕ i are depicted in Figure 3 (a) and the used measurement system is shown in Figure 3 (b). The measurement length in axial direction is l m = 29.7 mm. The contour at each angular position ϕ i is measured discretely and the distance between the measurement points is Δl = 0.5 μm. Thus, 59.400 measurement points are evaluated in axial direction. In general, the wear depth h W over the bearing width W is calculated by evaluating the difference between the contours before and after the experiment. At the angular measurement point ϕ i the wear depth h W (w j , ϕ i ) is determined for each axial measurement point w j with j t [0; 1; …; 59,400]. In Figure 3 (a) the wear depth at the front side of a journal bearing h W (w j = 0 mm, ϕ = 135°) is depicted. The planimetric wear (1) at the measurement point ϕ i can be approximated by summing up the products of the wear depth values h W (w j , ϕ i ) and the discrete axial measurement distance over the bearing width W. To approximate the wear volume (2) at the measurement point ϕ i the wear area A W (ϕ i ) is multiplied with the corresponding arc length Δϕ (ϕ i ). The arc length Δϕ (ϕ i ) starts at the middle angular position between the measurement point ϕ i and the previous measurement point ϕ i-1 . The arc length Δϕ (ϕ i ) ends at the middle angular position between ϕ i and the following measurement point ϕ i+1 . The arc length Δϕ (180°) is shown in Figure 3 (a). The total wear volume (3) is defined as the sum of the wear volume V W (ϕ i ) at each measurement point ϕ i . The calculation of the wear area A W is illustrated in Figure 4 (a) for the measurement ( ) = , , ( ) = ( ) ( ) = ( ) Science and Research 33 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 (a) (b) 1 measurement prism 2 journal bearing 3 probe arm 4 mearement system 1 2 3 4 90° 200° 270° 225° 180° 160° 135° arc length (180°) bearing width W ( = 0 , = 135°) axial measurement direction Figure 3: Specification of the journal bearing test specimen (a) and the measurement system (b) e R for different operating conditions two operating points (A and B) are defined. The specific pressure p̅ ̅ , sliding speed v s and target temperature T t of both operating points are listed in Table 3. Two test specimen combinations are tested for both operating points each. The two test specimen combinations are defined as i and ii. Thus, in total four test runs (A.i, A.ii, B.i, B.ii) are performed on the test rig. The measurement parameters and the results are summarized in Table 3. The total wear volume V W and the mean wear depth h W ¯¯ over the bearing width at the lowest angular position (ϕ = 180°) of the bearing are evaluated. Each test run takes 20 h and is divided into two sections with 10 h each. Wear volume V W and mean wear depth h W ¯¯ are measured after 10 h and after 20 h of testing. The results are plausible, as the test runs A with higher pressure compared to B lead to more wear. Additionally, less wear is measured in the second part of each test run compared to the first one due to running-in processes [9]. The test results can also be used to determine the friction energy density e R experimentally. F LEISCHER [10] defines (4) as the quotient of friction work W F and wear volume V W . The friction work = Science and Research 34 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 point ϕ i = 180° of an exemplary bearing. Figure 4 (b) visualizes the approximation of the wear volume V W (ϕ i ) based on the arc length Δϕ (ϕ i ) at ϕ i = 180°. In order to determine the wear volume measurement uncertainty, the measurement is performed three times for the same test specimen under the same conditions. The mean measurement error range of the wear volume yields ΔV W,error = 0.61 mm 3 . Therefore, the uncertainty of the wear volume measurement is quantified with ΔV W =± 0.3 mm 3 . Possible reasons for measurement errors may be imprecisions in adjusting the angular position (Δϕ = ± 1°) or the starting point of the axial contour measurement (Δl start = ± 0.1 mm). The same procedure for a tactile roughness measuring device leads to the uncertainty of the roughness measurement of ΔR = ± 10 %, which was also applied for the sensitivity analysis (Table 1). 4.3 Calculation of the friction energy density based on experimental results The test series is performed on the same component test rig that was used and described in [9,11]. The tested journal bearings are made from a copper-tin alloy CuSn12Ni2-C (produced in a continuous casting process), the nominal bearing diameter is D = 120 mm and the nominal bearing width is W = 30 mm. To determine contour [ μm] [mm] W (180°) time t=0h time t=10h wear depth W (a) W (180°) W (180°) arc length (180°) (b) Figure 4: Determination of the wear area (a) and approximation of the wear volume (b) Operating point A B Specific pressure [ ] 35 25 Sliding speed [ / ] 0.1 0.1 Target temperature [° ] 55 55 Test specimen i ii i Ii Bearing diameter [ ] 120.02 120.03 119.98 119.99 Shaft diameter [ ] 119.89 119.92 119.91 119.86 Bearing clearance [ ] 0.130 0.110 0.070 0.135 Surface roughness [ ] 0.34 2.28 2.75 2.72 Surface roughness [ ] 0.56 0.48 0.61 0.60 Measurement results A.i A.ii B.i B.ii = 10 [ ] 4.82 4.78 0.72 0.45 = 20 [ ] 0.14 0.73 0.21 0.17 = 10 [ ] 14.0 14.0 2.1 1.3 = 20 [ ] 0.9 1.4 0.8 0.7 Table 3: Test parameter for the experiments (5) is calculated based on normal force F N , sliding speed v S and time integration of the coefficient of friction μ(t). Normal force F N and sliding speed v S are known from the defined operating conditions for the wear experiments on the test rig. The coefficient of friction over time μ(t) can be calculated using C OULOMB ’s law of friction. The normal force F N is known and the friction force F F (t) can be determined by dividing the measured friction moment M F by the known bearing radius. The wear volume V W is measured according to section 4.2. However, the measurement of wear volume V W shows an uncertainty of ΔV W = ± 0.3 mm 3 (see section 4.2). The uncertainty of the wear volume measurements leads to an uncertainty in the calculation of e R . The experimentally quantified values of the friction energy density e R are shown in Figure 5 (a) for each test run. The uncertainty in the calculation of the friction energy density Δe R due to the wear volume measurement uncertainty is represented by error antennas. Considering the uncertainty of the friction energy Δe R density due to measurement uncertainties of the wear volume, a range from 1.2 · 10 15 J · m -3 to 3.1 · 10 16 J · m -3 can be measured for both operating points A and B. Alternatively, the range can be determined using the F LEISCHER diagram by evaluating e R depending on the friction shear stress τ R and the linear wear intensity I h [10]. The asperity friction shear stress (6) can be calculated by multiplying the applied asperity contact pressure p a by the coefficient of friction μ. The linear wear intensity (7) = ( ) = = is defined as quotient of the wear depth h W and the sliding distance s R . The sliding distance is known for a given sliding speed v s considering the test time t. The wear depth h W can be measured. The mean wear depth h W ¯¯ at the angular measurement point ϕ = 180° is used for the calculation. The resulting e R values for each test run according to the F LEISCHER diagram are depicted in Figure 5 (b). The maximum uncertainty Δe r in the F LEISCHER diagram ranges from 1.0 · 10 15 J · m -3 to 3.5 · 10 16 J · m -3 . The dimension of the range is comparable to the determination method based on the measurement uncertainties of the wear volume. The neglect of the local distribution of wear by analyzing the mean wear depth at a single angular measurement point instead of the total wear volume may be a cause of the slightly greater scatter. The relative distribution of the e r values between the individual test runs is qualitatively equal. As both methods of calculating the friction energy density e R are comparable, in terms of a worst-case scenario the larger range is used to determine the uncertainty of the wear simulation method. Thus, the friction energy range shown in Table 1 is updated to the new range of e R,min = 1.0 · 10 15 J · m -3 and e R,max = 3.5 · 10 16 J · m -3 . 4.4 Wear calculation uncertainty Due to uncertainties in the determination of model parameters the calculation of the wear volume using the presented wear simulation method is subject to uncertainties. This section presents the resulting uncertainty of the wear simulation method itself. The wear calculation uncertainty is quantified based on the maximum calculation error in terms of the wear volume and depth. To simulate maximum wear, the five most influential parameters according to the sensitivity analysis in section 4.1 Science and Research 35 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 Figure 5: resulting uncertainty in the calculation of the friction energy density Δe R due to the wear volume measurement uncertainty (a) and range of e R based on the Fleischer diagram (b) (a) (b) Taking Figure 6 into account, the maximum relative wear volume uncertainty for the considered operating points is quantified with ΔV W,rel < ± 49 %. The maximum relative wear depth uncertainty is Δh W,rel < ± 48 %. This very high uncertainty can be attributed to the high parameterization inaccuracy in the wear parameter. In reality it can be expected that the actual calculation error will be smaller, as this is a worst-case scenario. 5 Summary Unlike rolling-element bearings, journal bearings have a theoretically unlimited fatigue lifetime and are therefore particularly reliable when designed and operated correctly. However, external load conditions can lead to wear that might lead to seizure of the bearing. Thus, wear simulations are used to optimize the journal bearing design. Currently, wear simulations are used to predict the wear behavior. In order to determine the accuracy of the used method, the worst-case scenario calculation was performed in this work. First, the most significant parameters of the wear model were identified by applying a sensitivity analysis according to M ORRIS [12]. The highest influence was shown for the friction energy density e R . Thus, a precise determination of e R is required. However, the value is unknown a priori and dependent on the operating point. In addition, the parameter range from lite- Science and Research 36 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 are configurated to generate maximum wear. The configuration is based on the lower and upper parameter boundaries of the sensitivity analysis. In contrast to the one-factor-at-a-time approach, all parameters take the extreme value here. The remaining parameters are kept constant in the medium value range. Since the sensitivity analysis showed the greatest influence of e R on the wear volume, the minimum and maximum values from the experimental determination (section 4.3) are used instead of the boundary values from the literature to increase the accuracy. The parameter values for maximum and minimum wear simulations are summarized in Table 4. Each operating point with p̅ t [10 MPa; 15 MPa] and v s t [0.1 m · s -1 ; 0.2 m · s -1 ] is simulated for minimum and maximum wear for a duration of 10 h. The absolute uncertainty is defined as the difference between the minimum and maximum wear metric at the simulation end. Minimum wear volume V W,min , maximum wear volume V W,max and the absolute uncertainty of the wear volume ΔV W,abs for each operating point are shown in Figure 6 (a). Analogous, Figure 6 (b) visualizes the wear depth values h W and the uncertainties of the wear depth Δh W,abs . The absolute wear uncertainties increase for operating points with higher wear intensity. Thus, the absolute uncertainties of both wear metrics depend on the operating point. Both metrics show a comparable relative uncertainty, regardless of the operating point. Parameter (variated) Unit Max. wear Min. wear Parameter (constant) Unit Med. wear Elastic factor - 0.003 0.0003 Roughness orientation - 50 Friction energy density 1.0 10 15 3.5 10 16 Roughness parameter ± 0 % Model coefficient (O&K) - 1,000 10,000 Temperature ° 55 Model coefficient (O&K) - 1,000 50 Bearing clearance 0.15 Reference length 2 10 -6 1 10 -6 Coefficient of friction - 0.1 Table 4: Parameter values for maximum and minimum wear Figure 6: Quantified uncertainty of the wear calculation with regard to wear volume (a) and wear depth (b) (a) (b) rature is large. To realistically limit the parameter range, experimental tests were conducted on the component test rig. The minimum e R,min = 1.0 · 10 15 J · m -3 and maximum e R,max = 3.5 · 10 16 J · m -3 were determined for the considered operating points. The resulting absolute wear uncertainties ΔV W,abs and Δh W,abs are dependent on the operating point and increase for higher wear intensity. The relative wear uncertainties ΔV W,rel and Δh W,rel are similar and show a limit of < 49 % for the worst-case scenario. A more realistic uncertainty due to the imprecise determination of the model parameters may be significantly lower than this value. Also validation experiments of the presented method show, that valid wear simulation results can be achieved through precise parameter adjustments, when the real system behavior is well known [9]. The uncertainty identified in this work poses a challenge for the wear calculation, especially if no experimental confirmation tests can be carried out. 6 Outlook In this work it was demonstrated that the high measurement uncertainty in the determination of the wear volume V W jeopardizes the parameterization of the wear coefficient e R . To improve on the parameterization the determination of the wear volume should be further investigated in future work. In addition, for the two different phases of the wear experiments (test runs 1 and 2) and the two tested operating points different values for the wear coefficient e R were determined. To improve the parameterization of the wear model a time dependent wear coefficient could be introduced using the approach by L IJESH [21] as suggested by L EHMANN ET. AL. in [11]. Lastly, the used wear simulation model should be extended to include thermal effects. This has the potential to further increase the model fidelity and reduce deviations to the real-world wear behavior. References [1] Statistisches Bundesamt. 2024, “Stromerzeugung 2023: 56 % aus erneuerbaren Energieträgern,” https: / / www.destatis.de/ DE/ Presse/ Pressemitteilungen/ 20 24/ 03/ PD24_087_43312.html, accessed April 23, 2024 [2] Deutsche WindGuard. 2023, “Status of Onshore Wind Energy Development in Germany: Year 2023” [3] T. Thys and W. Smet. 2023, Selective assembly of planetary gear stages to improve load sharing [4] M. Prölß. 2020, Berechnung langsam laufender und hoch belasteter Gleitlager in Planetengetrieben unter Mischreibung, Verschleiß und Deformationen. Düren [5] T. Hagemann, H. Ding, E. Radtke and H. Schwarze. 2021, “Operating Behavior of Sliding Planet Gear Bearings for Wind Turbine Gearbox Applications - Part II: Impact of Structure Deformation,” Lubricants, Vol. 9, 98, DOI: 10.3390/ lubricants9100098 [6] H. Ding, Ü. Mermertas, T. Hagemann and H. Schwarze. 2024, “Calculation and Validation of Planet Gear Sliding Bearings for a Three-Stage Wind Turbine Gearbox,” Lubricants, Vol. 12, 95, DOI: 10.3390/ lubricants12030095 [7] F. König, C. Sous and G. Jacobs. 2021, “Numerical prediction of the frictional losses in sliding bearings during start-stop operation,” Friction, Vol. 9, 583-597, DOI: 10.1007/ s40544-020-0417-9 [8] F. König, A. Ouald Chaib, G. Jacobs and C. Sous. 2019, “A multiscale-approach for wear prediction in journal bearing systems - from wearing-in towards steady-state wear,” Wear, 426-427, 1203-1211, DOI: 10.1016/ j.wear.2019.01.036 [9] T. Decker, G. Jacobs, C. Graeske, J. Röder, M. Lucassen and B. Lehmann. 2024, “Multiscale-simulation method for the wear behaviour of planetary journal bearings in wind turbine gearboxes,” Journal of Physics: Conference Series, Vol. 2767, 52012, DOI: 10.1088/ 1742- 6596/ 2767/ 5/ 052012 [10] Gerd Fleischer. 1980, Verschleiß und Zuverlässigkeit. Berlin [11] B. Lehmann, P. Trompetter, F. G. Guzmán and G. Jacobs. 2023, “Evaluation of Wear Models for the Wear Calculation of Journal Bearings for Planetary Gears in Wind Turbines,” Lubricants, Vol. 11, 364, DOI: 10.3390/ lubricants11090364 [12] M. D. Morris. 1991, “Factorial Sampling Plans for Preliminary Computational Experiments,” Technometrics, Vol. 33, 161, DOI: 10.2307/ 1269043 [13] J. A. Greenwood and J. H. Tripp. 1970, “The Contact of Two Nominally Flat Rough Surfaces,” Proceedings of the Institution of Mechanical Engineers, Vol. 185, 625-633, DOI: 10.1243/ PIME_PROC_1970_185_069_02 [14] N. Patir and H. S. Cheng. 1978, “An Average Flow Model for Determining Effects of Three-Dimensional Roughness on Partial Hydrodynamic Lubrication,” Journal of Lubrication Technology, Vol. 100, 12-17, DOI: 10.1115/ 1.3453103 [15] G. Offner and O. Knaus. 2015, “A Generic Friction Model for Radial Slider Bearing Simulation Considering Elastic and Plastic Deformation,” Lubricants, Vol. 3, 522-538, DOI: 10.3390/ lubricants3030522 [16] F. König. 2020, Prognose des Verschleißverhaltens ölgeschmierter Gleitlager im Mischreibungsbetrieb. Aachen [17] AVL List GmbH. 2022, EXCITE Power Unit [18] M.-E. Kyrkou and P. G. Nikolakopoulos. 2020, “Simulation of thermo-hydrodynamic behavior of journal bearings, lubricating with commercial oils of different performance,” Simulation Modelling Practice and Theory, Vol. 104, 102128, DOI: 10.1016/ j.simpat.2020.102128 [19] C. Wolf. 2014, Reibungs- und Verschleißsimulation instationär belasteter Gleitlager unter elastohydrodynamischen Bedingungen. Kassel [20] A. Saltelli, S. Tarantola, F. Campolongo and M. Ratto. 2002, Sensitivity Analysis in Practice [21] K. P. Lijesh and M. M. Khonsari. 2018, “On the Modeling of Adhesive Wear with Consideration of Loading Sequence,” Tribology Letters, Vol. 66, DOI: 10.1007/ s11249- 018-1058-2 Science and Research 37 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0036 contrast, the dynamic tests consider the tribological load and are therefore well suited for detecting incompatibilities [6]. However, the disadvantages are the amount of lubricant required and the overall increased test costs. As part of the project FVA 578 III: ‘Tribologisches Ersatzsystem für dynamische Elastomer-Schmierstoff Verträglichkeitsprüfungen’, a new dynamic test methodology was investigated on a simple tribologically equivalent system for RSS, the ring cone tribometer (RFT). The aim was the development of a cost-effective dynamic compatibility test, which can already be used by interested companies at this time. In this article, the test development methodology is first presented. Furthermore, the results of the test runs on the RFT and the comparison of these with static and RSS tests are shown. Science and Research 38 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0037 1 Introduction In the lubricant development process the compatibility of a lubricant with rubber plays an important role. Incompatibility with rubber used for manufacturing rotary shaft seals (RSS), which are used to seal machines, can lead to unwanted leakage or even failure of the entire system during operation. Therefore, during the multistage lubricant development process, lubricant manufacturers carry out compatibility tests. If an incompatibility is only detected at a late stage, this means a high loss of investment and working time. Currently lubricant manufacturers have two types of compatibility tests at their disposal. These are on the one hand static tests (e.g. DIN ISO 1817), in which an elastomer sample is immerged in the test lubricant, and on the other hand dynamic tests (e.g. FLENDER and SEW tests), in which the compatibility is checked on RSS under real operating conditions [1], [2], [3]. The disadvantage of static tests is that they do not take tribological contact into account and therefore this test method does not detect incompatibilities reliably. In addition, the test conditions are only vaguely defined in some cases, meaning that good repeatability and comparability are not guaranteed [4], [5]. In A new rubber-lubricant compatibility test on a tribometer for radial shaft seals Laura Stubbe, Yvo Stiemcke, Sarah Mross, Sarah Staub, Konrad Steiner, Kerstin Münnemann, Oliver Koch, Stefan Thielen* submitted: 20.09.2024 accepted: 21.01.2025 (peer review) Presented at GfT Conference 2024 The present study deals with the development of a new dynamic test method to evaluate the elastomerlubricant compatibility on a tribologically equivalent system for radial shaft seals (RSS), the ring cone tribometer (RFT). The suitability of the RFT for compatibility testing was demonstrated based on tests on RSS and static aging tests with comparable operating conditions. There was good agreement between RSS and RFT tests, whereas static aging did not produce valid results due to the lack of tribological loading. The tests showed that the course of the friction coefficient and the optical analysis are highly informative regarding an assessment of compatibility. However, the occurrence of leakage, a change in hardness and the course of cell lowering during the test run also allow conclusions to be drawn about incompatibility. Keywords rubber, lubricant, compatibility, test method, tribometer, RSS, friction coefficient Abstract * Laura Stubbe, M.Sc. 1 (corresponding author) Dipl.-Ing. Yvo Stiemcke 1 Sarah Mross, M.Sc. 2 Prof. Dr.-Ing. Sarah Staub 3 Dr. Konrad Steiner 3 Dr. rer. nat. Kerstin Münnemann 2 Prof. Dr.-Ing. Oliver Koch 1 Jun. Prof. Dr.-Ing. Stefan Thielen 1 1 RPTU Kaiserslautern-Landau Chair of Machine Elements, Gears and Tribology (MEGT) Gottlieb-Daimler Str. 42, D-67661 Kaiserslautern Germany 2 RPTU Kaiserslautern-Landau Laboratory of Engineering Thermodynamics (LTD) Gottlieb-Daimler Str. 42, D-67661 Kaiserslautern Germany 3 Fraunhofer-Institut für Techno- und Wirtschaftsmathematik ITWM, Fraunhofer Platz 1, D-67663 Kaiserslautern Germany 2 Experimental Studies The RFT was developed as part of FVA 578 I and optimized in FVA 578 II with regard to the conformity of friction and wear with the RSS system [7], [8]. Figure 1 shows the schematic structure of a RFT test cell. The sealing contact is simulated with its air and oil-side contact angles α and β by the contact of a shaft cone with an elastomer ring sample. This has an inner diameter of 50 mm and an outer diameter of 75 mm. The rubber ring specimen is clamped in a specimen holder so that it cannot twist. This is axially movable and a line load equivalent to the radial force of the RSS can be applied to the ring specimen via weights. Above the sealing contact there is a temperature-controlled oil reservoir which can hold a lubricant quantity of approx. 190 ml. During the test run, the shaft cone is set in rotation and the frictional torque generated in the sealing contact can be recorded using a torque measuring hub. The wear of the ring sample can already be estimated during the test run (on-line) by means of eddy current sensors via the wear-related lowering of the sample holder [9]. As a result, the course of the change in the sensor-cell distance can be determined. A total of six rubber-lubricant combinations were tested, whereby each combination was tested twice. As rubbers, 75 FKM 585 and 72 NBR 902 were selected. For the lubricants, two practical lubricants (“High-Ref”) and two low-additivated base oils (“Low-Ref”) based on PAO or mineral oil were used, all with an ISO VG 220 viscosity. The shaft cones were made from 16MnCr5. Table 1 provides an overview of the resulting combinations and the corresponding test conditions. The oil sump temperature was set to 80 °C for combinations with NBR and 110 °C for combinations with FKM. The material-specific temperature was selected in such a way that chemical damage to the elastomer is avoided, but thermal acceleration according to A RRHENIUS takes place. Thereby, according to V AN’T H OFF , a temperature increase of 10 °C increases the reaction rate by a factor of two to three [11]. The test duration was 1008 h and the sliding speed was set to 6 m/ s in each case. Both parameters were determined experimentally. For this purpose, preliminary investigations were carried out in which the test duration (240, 1008 and 2016 h) and the speed collective were varied. It turned out that clear intolerance reactions can be identified after 1008 h and that the tests can be carried out at a stationary speed, which reduces the complexity of the tests compared to carrying them out with a speed cycle. The curves of the friction coefficient and the sensor-cell distance, the planimetric sealing edge wear and the change in hardness were evaluated. The determination of the values is described in the following. To determine the line load required for the ring specimens on the RFT, the radial force F r on the RSS heated Science and Research 39 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0037 Figure 1: Scheme of the ring cone tribometer test rig (RFT) and the contact angles based on [10] # Rubber Lubricant v slide / m/ s T Oil sump / °C Duration / h 1 75 FKM 585 High-Ref PAO 6 110 1008 2 75 FKM 585 Low-Ref Min 6 110 1008 3 75 FKM 585 Low-Ref PAO 6 110 1008 4 72 NBR 902 High-Ref Min 6 80 1008 5 72 NBR 902 Low-Ref Min 6 80 1008 6 72 NBR 902 Low-Ref PAO 6 80 1008 Table 1: Summary of test parameters IRHD hardness measurement was carried out based on the standard DIN ISO 48-2 [13] on the “digi test II” device from Bareiss ® , whereby method M (micro hardness) was selected. The hardness of the sealing edge of the ring samples and the RSS is measured before and after the tests at 8 points distributed around the circumference and the mean value is calculated from this. In addition, the analysis methods NMR, DSC, DMTA and optical analysis were evaluated, and a possible leakage was considered. These criteria were evaluated according to the evaluation scheme shown in Table 2. Values colored green indicate good compatibility and values colored red indicate incompatibility. Orange colored means that the result points to slight incompatibility. In addition to the tests on the RFT, static aging tests and RSS tests with comparable test conditions were carried out in order to compare the different methods with regard to the reliability of detecting an incompatibility. The RSS tests were carried out on the multi-shaft test bench [14]. RSS according to DIN 3760 A-80-100-10 were used. The static aging tests were carried out based on the standard DIN ISO 1817 [1]. 3 Results The test results of two elastomer-lubricant combinations are shown below. The results of a well-compatible combination are presented on the one hand and of an incompatible combination on the other. This should make it clear which criteria can be used to assess incompatibility in the RFTtests. No leakage occurred with the combination FKM - High-Ref PAO (#1), for which good compatibility was expected. The optical analysis shows no defects at the sealing edges. The friction coefficient curves show a steady course with few fluctuations within the first 400 h (Figure 3 (l)). Figure 4 shows the course of the sensor-cell distance. This shows a slight lowering, which Science and Research 40 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0037 to test temperature is first determined. The line load p l can be calculated from this and the diameter of the ring specimen D using the following formula. (1) The friction coefficient µ is calculated using the following formula from the friction torque M r measured during the test run, the sealing edge radius of the sample r and the line load p l . This eliminates the influence of the different specimen diameters on the RFT and the RSS system and the results can therefore be better compared with each other. (2) On the RFT, the wear of the ring sample can already be estimated during the test run (on-line) by means of eddy current sensors via the wear-related lowering of the sample holder [9]. This results in the course of the cell lowering during the test run. After the test run, the planimetric wear is determined as described in [12]. The micro Criteria Evaluation Leakage / ml 0 > 0; ≤ 5 > 5 Friciton coefficient curve Slight changes / Stationary course Unsteady course without fluctuations Strong fluctuations Change of sensor-cell distance* Slight swelling, low wear Minor fluctuations in the course Significant fluctuations in the course Increased wear Planimetric wear / mm 2 < 0,005 0,005 < x < 0,01 > 0,01 Hardness change / IRHD-M <+2; >-6 <+5 ; >-10 > +5 / < -10 Optical analysis Inconspicuously Discoloration Cracks, blistering, chemical erosion Table 2: Evaluation scheme for assessing the analytical methods regarding compatibility. Figure 2: Optical analysis of the sealing edge of the samples for the combination NBR - Low-Ref Min.. L.: Blistering and discoloration (RFT), m.: Cracks and discoloration (RFT), r.: Blistering and discoloration (RSS). * The investigated combinations did not show a critical case of swelling, therefore no red category was defined. indicates low wear, and also a partial raising of the test cells, which indicates swelling. The results of the planimetric wear measurement confirm the moderate wear expected from the course of the sensor-cell distance (Figure 5). The change in hardness (Figure 6 (l)) shows a clear decrease in hardness for all ring samples, which indicates swelling. This is confirmed by the NMR, as shown in Figure 7. The NMR measurements of the samples (for details see Ref. [10]) show a clear lubricant signal at the sealing edge. The resolution of the NMR measurement of ~0.16 +-0.03 mm/ pt allows a rough determination of the penetration depth based on the length of the signal peak. This results in a penetration depth of approx. 0.7 mm. Together with the reduced hardness and the decrease in distance in the measurement of the sensor-cell distance, the test results confirm a slight swelling in the FKM - High-Ref PAO combination. However, this does not appear to affect the sealing function. As with all analytical methods, it is therefore necessary to determine limit values from which a change is classified as negative. For the combination NBR - Low-Ref Mineral oil (#5), for which an incompatibility was expected, significant leakage was detected. All ring samples show clear discoloration, cracks and, in some cases, blistering of the sealing edge (Figure 2 (l),(m)). The friction coefficient curves of the first 400 h are shown in Figure 3 (m). The tests show a strong change in the friction coefficient, especially at the beginning of the test run. This is an indicator of incompatibility [15], [16]. The progression of the sensor- Science and Research 41 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0037 Figure 3: Friction coefficient curves of the first 400 h. L.: FKM - High-Ref PAO on RFT (on 3 samples), m.: NBR - Low-Ref Min. on RFT (on 3 samples), r.: NBR - Low-Ref Min on RSS (on 2 samples). Figure 4: Course of sensor-cell distance for the combinations FKM - High-Ref PAO and NBR - Low-Ref Min. on RFT. During some test runs, the data acquisition system failed, which is why no curve can be shown for this time period. Figure 5: Planimetric wear comparison of the combinations FKM - High-Ref PAO and NBR - Low-Ref Min. (average of three values). ples of the combination NBR - Low-Ref Min. (#5) showed a slight shift of T g by 1.3 °C or 2 °C in a positive direction. Due to the low informative value of the DSC with regard to the evaluation of compatibility in the RFT tests, this analysis method is omitted from further consideration, just like the DMTA. In total, the evaluation of the RFT test runs showed clear incompatibility reactions for selected combinations and criteria, as shown in Table 3. According to the results, for combination 3-6 premature failure would be expected during usage. The criteria “Optical analysis” and “Friction coefficient curve” in particular can be used to draw reliable conclusions about incompatibility in the RFT tests. Thereby significant fluctuations in the friction coefficient curve indicate incompatibility (compare Figure 3), as also shown in [15]. When observing the change in the sealing edge, clear cracks, blistering and discoloration were found in the incompatible combinations 4-6 (compare Figure 2), and chemical erosion of the sealing edge in combination 3. The comparison of the different test methods regarding the evaluation of the compatibility of the combinations shows a good agreement between the RFT tests and the RSS tests, as shown in Tables 3 and 5. With the incompatible combination NBR - Low-Ref Min., the RSS tests also show clear fluctuations in the friction coefficient curve (Figure 3 (r)) as well as discoloration and blistering on the sealing edge (Figure 2 (r)). However, it was shown that incompatibility reactions are probably provoked more quickly on the Science and Research 42 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0037 cell distance shows a clear increase in the axial position of the test cells, which indicates increased wear (Figure 4). This is confirmed by the measurement of the planimetric wear (Figure 5). It should be noted that the measurement was falsified by the damage to the ring samples (cracks, blistering) so that the absolute values are not valid for this combination. The hardness measurement shows a slight decrease in hardness for the samples on average (Figure 6 (r)). In the signal of the NMR analysis no change can be seen for this combination, which indicates that no lubricant has penetrated the sample. The DMTA measurement shows a clear decrease in the storage modulus of the FKM-ring sample and a shift in the peak of the loss factor (shift in T g ). The NBR-ring sample shows a decrease in the storage modulus, a shift in T g is not recognizable here. Although the DMTA showed changes in the elastomers, no causality between the changes and an incompatibility could be determined here due to the small amount of data. The results of the DSC are unremarkable for most samples, only the sam- Figure 6: Comparison of the change of IRHD hardness throughout the investigation for the combinations FKM - High-Ref PAO (l) and NBR - Low-Ref Min. (r) (average of three values). Figure 7: Result of the NMR for a sample of the combination FKM - High-Ref PAO. RFT than with RSS tests. In contrast, there is less agreement with the results of the static aging test (Table 5). This does not reliably identify incompatibility, as the tribological load is missing here. The results were used to develop a test recommendation for carrying out compatibility tests on the RFT. The criteria were evaluated regarding their suitability for assessing incompatibility. It was found that the friction coefficient curve and the optical analysis in particular provide reliable results. 4 Conclusion and Outlook As part of the project FVA 578 III [17], a new type of compatibility test was developed on the RFT, which reliably reveals incompatibilities. For this purpose, a test cycle and the test duration required to reliably detect incompatibility reactions were first determined using tests on the RFT. This is 1008 h and the tests can be carried out at a stationary speed. Afterwards, all test-relevant combinations were tested with the developed test parameters. The evaluation showed some clear signs of incompatibility, such as cracks and blistering formation at the sealing edge. Furthermore, clear reductions in hardness due to the tribological load could be determined, which, in combination with the course of the sensor-cell distance, indicated swelling of the corresponding samples. This assumption was finally confirmed by NMR analysis. The course of the friction coefficient was identified as another reliable indicator for recognizing incompatibility. Significant fluctuations in this curve indicate incompatibility. Overall, incompatibility between lubricant and elastomer in dynamic tribological contact on the RFT can already be simulated in early phases of lubricant development with laboratory-scale samples. By testing with tribological load, incompatibilities are detected much more reliably than with static tests. This has been shown by studies with comparable test parameters. The good comparability of the results with tests on RSS was also demonstrated based on the tests carried out. Another advantage of testing on the RFT is that it is cost-effective. As part of the FVA projects 578 I/ II [7], [8], the design of the RFT was optimized for simple production, mainly from standard parts. Interested companies were provided with all the necessary design drawings as well as a manual for setting up, carrying out tests, sample production and measurement. Due to the compact design of the RFT, it allows a very space-saving, cost-effective use of several test cells with only one drive. The lubricant requirement is significantly lower than for comparable RSS test rigs and the preparation of elastomer ring specimens from readily available test plates is very cost-efficient. Incompatibilities between elastomer and lubricant can be easily and reliably identified after the test by observing the friction coefficient curve and optical analysis of the test specimens. The changes in hardness and the course of the sensor-cell distance can also be used to draw conclusions about changes in the elastomer, in particular Science and Research 43 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0037 Table 3: Results of the tests on the RFT, evaluated in terms of compatibility. Green colored: Combination is well compatible, orange colored: Slight incompatibility reaction, red colored: Combination is incompatible Table 4: Results of the tests on RSS, evaluated in terms of compatibility. Green colored: Combination is well compatible, orange colored: Slight incompatibility reaction, red colored: Combination is incompatible Table 5: Results of the static tests, evaluated in terms of compatibility. Green colored: Combination is well compatible, orange colored: Slight incompatibility reaction Combination-# Leakage Friction coefficient curve Change in hardness Optical analysis NMR Compatible? 1 Yes 2 Yes 3 No 4 No 5 No 6 No Combination-# Leakage Friction coefficient curve Change in hardness Optical analysis NMR Compatible? 1 Yes 2 Yes 3 No 4 No 5 No 6 No Combination-# Leakage Friction coefficient curve Change in hardness Optical analysis NMR Compatible? 1 Yes 2 Yes 3 Yes 4 Yes 5 Yes 6 Yes Elastomeren für Dichtungsanwendungen in der Antriebstechnik,“ Forschungsvereinigung Antriebstechnik e. V. (FVA), Forschungsvorhaben 578 I, Abschlussbericht, FVA-Heft Nr. 979, 2011. [8] D. Weyrich, „Praxistaugliche Prüfmethodik für Reibungs- und Verschleißuntersuchungen am tribologischen Ersatzsystem von RWDR,“ Forschungsvereinigung Antriebstechnik e. V. (FVA), Forschungsvorhaben 578 II, Abschlussbericht, FVA-Heft Nr. 1374, 2020. [9] C. Burkhart, T. Schollmayer, B. van der Vorst, M. Sansalone, S. Thielen und B. Sauer, „Development of an online-wear-measurement for elastomer materials in a tribologically equivalent system for radial shaft seals,“ Wear, p. 203671, 2021. [10] D. Bellaire, S. Thielen, C. Burkhart, K. Münnemann, H. Hasse und B. Sauer, „Investigation of Radial Shaft Seal Swelling Using a Special Tribometer and Magnetic Resonance Imaging,“ ACS Omega 7/ 14, pp. 11671-11677, 2022. [11] S. Pongratz, „Alterung von Kunststoffen während der Verarbeitung und im Gebrauch,“ Universität Erlangen-Nürnberg, 2000. [12] T. Schollmayer, C. Burkhart, W. Kassem, S. Thielen und B. Sauer, „Verschleißanalyse an Radialwellendichtringen und weiteren Maschinenelementen mittels Laserprofilometrie,“ in Gesellschaft für Tribologie e. V. (Hg.): 62. Tribologie-Fachtagung, Online-Konferenz, 2021. [13] DIN Deutsches Institut für Normung e. V., DIN ISO 48- 2: Elastomere oder thermoplastische Elastomere - Bestimmung der Härte, Berlin: Beuth Verlag, 2021. [14] S. Thielen, P. Breuninger, H. Hotz, C. Burkhart, T. Schollmayer, B. Sauer, S. Antonyuk, B. Kirsch und J. C. Aurich, Improving the tribological properties of radial shaft seal countersurfaces using experimental micro peening and classical shot peening process, Tribology International, 2021. [15] C. Wilbs, D. Frölich, M. Adler und A. Heinl, „Radial lip seal friction torque - a suitable lubricant-elastomer compatibility indicator? ,“ in Nextlub, Düsseldorf, 2023. [16] A. Hüttinger, M. Woeppermann und J. Hermes, „Dynamische RWDR Tests neu definiert! ,“ in 60. Tribologie-Fachtagung, Göttingen, 2019. [17] L. Stubbe, „Tribologisches Ersatzsystem für dynamische Elastomer-Schmierstoff-Verträglichkeitsprüfungen,“ Forschungsvereinigung Antriebstechnik e. V. (FVA), Forschungsvorhaben 578 III, Abschlussbericht, 2024. Science and Research 44 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0037 swelling can be detected. This can be confirmed using NMR. The knowledge gained was used to develop a test recommendation for carrying out compatibility tests on the RFT. 5 Acknowledgement This work was funded by the German Federal Ministry of Economics and Climate Protection within the framework of Industrial Collective Research (IGF 21650N). The authors would like to thank the Forschungsvereinigung Antriebstechnik (FVA) for their support within the framework of the research project FVA 578 III “Dynamische Elastomer-Schmierstoffverträglichkeit”. References [1] DIN Deutsches Institut für Normung e. V., DIN ISO 1817: 2015: Elastomere oder thermoplastische Elastomere - Bestimmung des Verhaltens gegenüber Flüssigkeiten, Berlin: Beuth Verlag, 2016. [2] Freudenberg Sealing Technologies, Dynamic oil compatibility tests for Freudenberg radial shaft seals to release the usage in FLENDER-gear units applications (Table T 7300), 2018. [3] SEW Eurodrive, Prüfvorschrift 97 118 03 15: Statische und dynamische Prüfungen von Radialwellendichtringen (RWDR), 2016. [4] J. Braun, „Elastomerverträglichkeitsunterscuhungen von Schmierstoffen - Reicht die bestehende Normung aus,“ Tribologie + Schmierungstechnik, Bd. 56, Nr. 6, pp. 24- 29, 2009. [5] J. Braun, „Aktuelle Elastomerverträglichkeitstest-Standards - eine kritische Betrachtung,“ in Proceedings of the 16th International Sealing Conference, Stuttgart, 2010. [6] M. Klaiber, „Additivverträglichkeit: Einfluss verschiedener Additive auf Elastomere und die tribologischen Eigenschaften im System Radial-Wellendichtung,“ Forschungskuratorium Maschinenbau e.V. (FKM), Forschungsvorhaben 290, IGF-Nr.: 15903 N, Abschlussbericht, Heft Nr. 318, 2012. [7] T. Gastauer, „Vergleichende Reibungs- und Verschleißuntersuchungen durch Experimente und Simulation an Introduction and motivation For numerous activities in or with soils, the A BR value in accordance with NF P18-579 [1] is required to characterise the abrasiveness of the soil (cf. among others DIN 18312 [2]). The LCPC test required for this involves the use of a soil sample with a grain size distribution curve from 4.0 mm to 6.3 mm, which corresponds to the grain fraction of fine gravel. A steel disc (test wing) rotates in this sample for 5 minutes. The mass loss of the test wing is measured and the abrasiveness, the so called A BR value, is calculated. It is given in gram (abraded) material per tonne of soil material (g/ t). The standard specifies that the test wings must be made of rolled unalloyed quality steel C15 (1.0401), have a hardness between 60 and 75 HRB (Rockwell-B) and their surfaces must be descaled by sandblasting [1]. Furthermore, in [3] it is recommended that the surface of the test wings must be metallically bright and free of adhesions, in particular oxide layers, and that the hardness of the test wings must be verified by random batch testing. In addition to the grain size described above, investigations can be carried out with soils that have a grain size between 0.0 mm and 6.3 mm (cf. [4]) or 2.0 mm to 8.0 mm (cf. [5]). Previous work (including [6], [7], [8] in particular) do not take into account the material pro- Science and Research 45 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 Materials Investigation of Test Wings from LCPC Abrasiveness Test with Different Soils and Varying Test Duration Danka Katrakova-Krüger, Jonas Kotscha, Christoph Budach, Peter Erdmann* submitted: 20.09.2024 accepted: 24.01.2025 (peer review) Presented at GfT Conference 2024 For numerous activities in or with soils, information on the abrasiveness of the soil is required, as this is an important parameter for the economic efficiency of a construction project. To determine the abrasiveness of loose soil, the LCPC test is carried out in accordance with the French standard NF P 18-579. In this test, a test wing rotates for five minutes in a soil sample with a grain size between 4.0 mm and 6.3 mm. The mass loss is determined, from which the A BR value is calculated. The test wings used for the tests must have defined material properties so as not to falsify the abrasiveness results. This article presents material characterization of test wings from different test series using the LCPC test. In contrast to the French standard, the test wings are not only examined with regard to their mass loss, but their hardness is measured and a metallographic characterization of the metal microstructure is also carried out. In the tests, the grain fraction, the test duration and the fraction of broken grains in the soil sample were specifically varied. On the one hand, the metallographic investigations support and confirm the geotechnical results, on the other hand, they provide important insights into the wear of the test wings in the LCPC test. In the series of tests with a grain fraction that deviated from the standard, the influence of different grain sizes on the metal structure was also clearly demonstrated. The progression of wear on the test wings over time can be visualised in the tests with different test times. In addition, differences in the microstructure and surface properties of test wings, which can affect the resulting A BR values, were identified and analysed. Recommendations for the manufacturing of the test wings for the standardised LCPC test are derived. Keywords LCPC-test, abrasiveness, test wings, metallography, soil, soil grain size distribution curve, test duration Abstract * Prof. Dr. Danka Katrakova-Krüger (corresponding author) Jonas Kotscha M.Sc. Faculty of Computer Science and Engineering Science, Institute for General Mechanical Engineering, Materials Laboratory, University of Applied Sciences (TH) Cologne, Gummersbach, Germany Prof. Dr. Christoph Budach Faculty of Civil Engineering and Environmental Technology, Geotechnical Engineering and Tunneling, University of Applied Sciences (TH) Cologne, Cologne, Germany Prof. Dr. Peter Erdmann Faculty of Process Engineering, Energy and Mechanical Systems, Cologne Institute of Construction Machinery and Agricultural Engineering, University of Applied Sciences (TH) Cologne, Cologne, Germany As part of this study, the grain size distribution and the proportion of broken grains in the soil as well as the duration of the test were systematically varied. In addition to investigating the mass loss of the test wings, their hardness was determined and also a detailed characterization of the metal microstructure was also carried out. First results of these investigations - in particular the geotechnical aspects - were shown in [14] and [15]. In this contribution the focus is set on the microstructure investigation of the test wings in relation to the observed weight loss and the respective A BR value. Materials and methods Three test campaigns were carried out, in which the standardised parameters of the LCPC test were specifically varied. In test series A1 and A2 (test campaign A), the test duration was changed for two different soils. Science and Research 46 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 perties mentioned in [1] and [5] or have focussed on the properties of the test wings without specific variation of the grain size of coarse-grained soils. Also, the development of the wear progression over the time was not considered in combination with materials characterization. Other studies, including [9], [10], [11], [12] and [13], deal with LCPC tests under variation of soil parameters ( [9] and [10] ) as well as with the adjustment of test parameters, such as rotation speed [12]. In [13], a correlation analysis is carried out between the abrasiveness, hardness and mechanical properties of rocks exposed to different freeze-thaw cycles. In addition, [11] provides an overview of various laboratory tests for assessing abrasiveness. In contrast to these studies, the present work includes a detailed material characterisation of each test wing in addition to geotechnical analyses in order to provide a more comprehensive understanding of the influencing factors. Table 1: Boundary conditions and results of the LCPC tests carried out based on [16], [17] and [18] Figure 1: A BR value as a function of the test duration 0 - 60 minutes and the grain fraction Figure 2: A BR value as a function of the test duration 0 - 10 minutes and the grain fraction The aim of test campaign B was to investigate the influence of different grain fractions, i.e. grains of different sizes in the soil samples. The boundary conditions and results of the LCPC tests of test campaigns A and B as well as the hardness values of the test wings are shown in Table 1. The sample A1.5 was tested according to the standard LCPC conditions. In Figure 1 the abrasiveness results of the test series A1 and A2 are visualized. The A BR value increases as a function of the test duration. At the beginning, in the first ten minutes, the A BR value increases significantly more than later. Figure 2 shows an enlarged version of the curves for the first ten minutes. From the higher slope of the curves in the first ten minutes, especially the first five minutes, it can be concluded that the wear process is strongest at the beginning. From the tenth minute onwards, the wear rate decreases slightly and remains more or less constant for the rest of the time. Comparable results were obtained in the studies [6] and [7]. It can also be seen that the abrasiveness is higher for soils containing coarser grains (series A2). The results of test campaign C are listed in Table 2. In this test series, the fraction of broken grains in the soil sample was gradually increased by 20 % (C0 0 % - C5 100 %). Sample C0 serves as a reference sample and in this case corresponds to sample A1.5. Both designations refer to one and the same sample. The diagram (Figure 3) shows the A BR value as a function of the fraction of broken grains. It can be seen from the graph that the A BR value increases almost linearly with increasing fracture grain content in the soil sample. After completion of the LCPC tests and the geotechnical evaluation, the metallographic investigation was carried out using the following methodology. First of all, the test wings were visually inspected, their appearance described, and any special features documented. The hardness according to Rockwell B (HRB) was then recorded for all test wings. Three measurements were carried out in each case, from which the average value was determined. The hardness measurements were carried out after the LCPC tests. In order to avoid the influence of the test, the hardness indentations were placed in the area of the clamping of the test wing (cf. Figure 4). This area is unaffected by collisions with the soil grains during the LCPC test, which is why the material is expected to remain unchanged there and can be used for the hardness measurements on the wing material. Science and Research 47 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 Table 2: Boundary conditions and results of the LCPC tests carried out based on [18] Figure 3: A BR -value as a function of the proportion of broken grains including linear regression 10 minutes). In addition, tests with significantly longer test durations of 20, 40 and 60 minutes were carried out in test series A1. The progress of wear is shown in Figure 6 (for the shorter and the regular 5-minutes test duration) and in Figure 7 (for longer test duration). On the test wings, which were tested less than the regular 5 min, it can be observed that the material is initially compressed and strongly deformed. At both corners, areas form in which the test wing material is “smeared” over the edge. In Figure 8 and Figure 9 the samples with a test duration of one minute for both investigated soil grain fractions are given with more detail. With longer test durations, such as four minutes, these areas are increasingly eroded (see Figure 10 and Figure 11 at the upper edge). The heavily deformed areas lose their connection to the base material and become smaller. In addition, when comparing the samples from test series A1 and A2, it can be seen that the samples from series A2 wear faster. The heavily deformed areas become smaller more quickly than in the samples from test series A1. The reason for this lies in the soil sample used in each case. The soil sample from test series A2 Science and Research 48 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 To create micrographs, one of the opposing, heavily worn corners of the test wing was cut out and hot-embedded. Figure 5 shows how the test wing was cut. The small triangular piece at the bottom right of the image has been used. The sample is embedded so that the longest side (diagonal section / red line in Figure 5) can be observed and metallographically evaluated. After hot embedding, the sample was ground and polished in several passes. The finest polishing agent had a grain size of 1 µm. The sample was etched in three per cent nitric acid. The sample preparation and analysis correspond to the one described in [9]. The test wing shown in Figure 4 and Figure 5 has been stored in the normal laboratory environment after the LCPC test, which is why it has a slight oxidised layer. Results and discussion Variation of the test duration In test campaign A, various LCPC tests were carried out with different test durations for the grain fractions 4.0 mm - 6.3 mm and 2.0 mm - 8.0 mm. The test duration varied from the standardised five-minute test duration, being shorter (1, 2, 3 and 4 minutes) and longer (7.5 and Figure 4: Test wing after hardness test Figure 6: Wear progress LCPC test duration 1, 2, 3, 4 and 5 minutes (from left to right) grain fraction 4.0 mm - 6.3 mm Figure 5: Test wing separated has larger maximum grain size (8.0 mm) than that from test series A1 (6.3 mm), which is in accordance with the determined higher A BR values (cf. Table 1 and Figure 2). When analysing sample A1.60, which was tested for 60 minutes in the LCPC test, it can be clearly seen that the entire sample is very heavily worn, and a lot of material has been removed. In the area of the left edge, near Science and Research 49 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 Figure 7: Wear progress LCPC test duration 7.5, 20 and 60 minutes (from left to right) grain fraction 4.0 mm - 6.3 mm Figure 8: Sample A1.1 Test duration 1 minute grain fraction 4.0 mm - 6.3 mm Figure 9: Sample A2.1 Test duration 1 minute grain fraction 2.0 mm - 8.0 mm Figure 10: Sample A1.4 Test duration 4 minutes grain fraction 4.0 - 6.3 mm Figure 11: Sample A2.4 Test duration 4 minutes grain fraction 2.0 - 8.0 mm produce a large deformation layer on the test wing. This phenomenon is investigated and described in more detail in test campaign B. The tests with varying test durations are needed to validate a numerical simulation. The LCPC test was digitally modelled. A discrete element method (DEM) simulation of the LCPC test was carried out using the EDEM software. The digital replacement model is shown in Figure 15. When evaluating the simulation results, both the wear pattern on the test blade and the overall wear progress (cf. Figure 6 and Figure 7) as well as the changes in the soil sample (see Figure 15) can be analysed and compared. The experimentally obtained test results, including the micrographs and soil samples, serve as a reference. Science and Research 50 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 the outermost point in Figure 12 the test wing has a thickness of approx. 2.5 mm. Compared to the unworn sample, half of the material thickness has been removed. Overall, the test wing has taken on an aerodynamic shape. Sample A1.60 also has a different surface which is much finer and softer to the touch. A change in the deformed surface layer can be seen in the micrograph. The surface layer of sample A1.60 is approx. 19.37 µm thick (cf. Figure 13). If the surface layers of samples A1.1 to A1.60 are compared, it is noticeable that the deformation layer generally decreases over the test duration. The surface layer is increasingly removed as the LCPC test progresses. Furthermore, apparently no new surface layer is built up. This can be attributed to the fact that the bigger grains of a soil sample are crushed during the LCPC test (cf. Figure 14) and the remaining smaller grains do not Figure 12: Sample A1.60 Test duration 60 minutes grain fraction 4.0 mm - 6.3 mm Figure 13: Surface layer sample A1.60 grain fraction 4.0 mm - 6.3 mm Figure 14: Example soil sample before (left) and after (right) the LCPC test [19] Variation of the grain size distribution of the soil samples In test campaign B, a soil comparable to test campaign A in terms of its mineral composition was used for the LCPC trials. In the tests of test campaign B, the grain size distribution curve of the soils was deliberately varied further by also taking sand fractions into account. In principle, the results of the geotechnical investigations can also be observed and confirmed here metallographically. The increased wear of samples that were tested in a larger grain fraction (larger maximum soil grain size) can be clearly recognised on the micrographs. This can be seen in the comparison of sample B1 (Figure 16) and sample B2 (Figure 17). Sample B2 is significantly more rounded over the entire height of the test wing. With sample B1, rounding can only be recognised in the area of the corners. Further conclusions can be drawn from the micrographs. Due to the mechanical impact of the soil grains on the test wing during the LCPC test, cold deformation occurs on the surface of the test wing. Deformed material can be seen in the edge area of the heavily worn sample B2. The crystallites near the surface show a clear compression, in fact the material has been work-hardened (see Figure 19). Looking at the only slightly worn sample B1 (Figure 18), it is noticeable that the deformed edge layer is not so well pronounced here, or cannot be recognised at all. This finding also confirms that soil samples with a smaller maximum grain size (grain fraction) are less abrasive than soil samples with a larger grain fraction (cf. among others [15]). Further findings from test campaign B relate to the production of the test wings. When analysing the micrographs of sample B4, special features emerged that were not observed in the other samples of both test campaigns. Science and Research 51 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 Figure 15: Digital replacement model of the LCPC test [20] Figure 16: Micrograph sample B1 grain fraction 0.063 mm - 2.0 mm after 5 minutes - A BR = 468 g/ t Figure 17: Micrograph sample B2 grain fraction 0.25 mm - 6.3 mm after 5 minutes - A BR = 1138 g/ t wing was done in an oxygen-containing atmosphere. Normally, nitrogen atmosphere or vacuum are used for heat treatment in order to prevent possible surface layer decarburisation. The extent to which these deviating microstructural findings influence the results of the LCPC test was not investigated in depth yet. Taking into account the findings from [8] it can be assumed that the wear in the decarburised areas will be significantly higher, as there are almost no hard phases (pearlite, containing cementite Fe 3 C) in the microstructure, but only soft ferrite. This could not be suspected when looking only at the material hardness which in this case is within the specified range and also comparable with the other tested wings. But the comparability of the abrasiveness results is no longer given, at least from a material point of view. Based on [8] and on the results of own investi- Science and Research 52 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 When comparing sample B4 (Figure 20 and Figure 21) with the samples B1 and B2 (Figure 16 and Figure 17), it is noticeable that the perlite lines of sample B4 run vertically and not horizontally, i.e. the rolling direction is different. This leads to the conclusion that sample B4 was manufactured from a different semi-finished product. On the other hand, there is an approx. 0.5 mm thick layer in the upper and lower area, which is different. The microstructure there is considerably coarser-grained than the rest of the sample. This might be due to longer time or higher temperature during the heat treatment. In addition, almost no perlite can be recognised. This means that the carbon content is much lower in those areas. All these observations point to surface layer decarburisation during the annealing process. Surface layer decarburisation can occur if the heat treatment of the test Figure 18: Micrograph sample B1 enlarged 500 times grain fraction 0.063 mm - 2.0 mm Figure 19: Micrograph sample B2 enlarged 500 times grain fraction 0.25 mm - 6.3 mm Figure 20: Micrograph sample B4 grain fraction 0.063 mm - 8.0 mm Figure 21: Micrograph sample B4 enlarged 50 times grain fraction 0.063 mm - 8.0 mm gations, it is recommended to include the production processes (same semi-finished product, same rolling direction, cooled CNC milling, heat treatment in inert gas atmosphere or vacuum) of the test wings in the standard of the LCPC test [1] or the recommendation [3] and clearly define them. It is also advisable to specify the storage conditions of the test wings to prevent corrosion on the surface. This can be ensured by lightly oiling the surface and storing in a dry place. Random micrographs of the test wings before they are used for abrasiveness testing should be taken to detect possible deviation in the microstructure which might be a reason for misleading results as this cannot be predicted by the hardness alone. Change in the fraction of broken grains In test campaign C (change in fraction of broken grains), the geotechnical observation of increasing abrasiveness is difficult to confirm in terms of material characterisation. The purely visual inspection of the micrographs of samples C1, C3 and C5 (Figure 22, Figure 23 and Figure 24) makes it difficult to prove the increasing abrasiveness. For a clear confirmation of the geotechnical results, the increase in abrasiveness due to the increasing fraction of broken grains in this test campaign is too small. Overall, the micrographs show similar findings. Science and Research 53 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 Figure 22: Micrograph of sample C1 Figure 23: Micrograph of sample C3 Figure 24: Micrograph of sample C5 Figure 25: Thickness of the surface layer test campaign C ! "#$"# % $' ' $#! #? $%! ? $ ' Table 3: Thickness of the surface layer test campaign C information on the wear progression of the test wings. The tests show that the test duration and the grain fraction have a significant influence on wear. After an initial high abrasion, the wear rate diminishes and remains nearly constant. Higher wear is observed for soils with higher maximum grain size. Also, an influence was found on the deformation layer built on the surface during the test. With increasing proportion of broken grains, the abrasivity of the soil is increasing. The thickness of the deformation layer seems to follow correspondingly but is significantly lower for 100 % broken grains. The reason for this is not ultimately clarified yet and further investigation is needed. Significant differences were found in the surface quality and the microstructure of the test wings, which can affect the comparability of the A BR values. Based on these findings, recommendations were derived for the production of standardised test wings and their storage in order to achieve more reliable results on the abrasiveness of soils. Currently, a round robin test is being carried out in which the LCPC tests of around 20 laboratories and testing facilities are being compared. The laboratories are provided with controlled soil samples and test wings for the LCPC tests. In addition, the participating labs are using own test wings for the LCPC tests. The results of the round robin test can be used to evaluate the deviation between the test devices as well as the scattering of the LCPC tests with different test wings. Also, effort is put on the Discrete Element Method (DEM) simulation of the abrasiveness test to be able to understand the phenomena as well as to predict results and significantly reduce experimental costs. Acknowledgements This work was supported by the Ministry of Culture and Science of North-Rhine Westphalia, Germany within the program HAW Kooperation. Science and Research 54 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 The analysis of the surface layer shows a special feature. Basically, it was observed that a deformed and presumably work-hardened surface layer forms on the test wings. This surface layer initially increases proportionally with the increasing fraction of broken grains. The measured values are listed and illustrated in Table 3 and Figure 25. This development runs parallel to the increase in abrasiveness (A BR value). However, sample C5 represents an anomaly. Whereas the thickness of the surface layer increases continuously in samples C0 to C4, it drops suddenly and rapidly in sample C5. The sharply decreasing thickness of the surface layer from sample C4 to C5 is shown in Figure 26 (C4) and Figure 27 (C5). The origin of this deviation is not finally clarified up to now. One possible reason is that the deformed layer may have been more pronounced but already abraded. To check this, tests with varying time with this type of soil must be performed. Further investigations into this are planned but have not yet been realised as the compilation of a defined soil sample with an exact fraction of broken grains is very time-consuming. The major time expenditure is in the manual sorting and characterisation of the soil sample’s aggregates which is also dependent on the person doing this sorting. Summary and outlook This paper describes materials characterization of test wings after investigations of the abrasiveness of soil using the LCPC test according to the French standard NF P 18-579. The LCPC test determines the mass loss of a rotating steel test wing in a soil sample in order to calculate the A BR value. In the test campaigns carried out, the test durations and the soil grain size distribution curves were varied in order to investigate their influence on the abrasiveness and wear of the test wings. The results of the metallographic analyses fit well to the geotechnical results and provide Figure 26: Thickness of the surface layer sample C4 Figure 27: Thickness of the surface layer sample C5 Remarque Part of this research was presented at the 65th German Tribology Conference 2024 taking place 23.-25-09.2024 in Göttingen, Germany [21]. References [1] NF P 18-579, Gesteinskörnungen - Bestimmung der Koeffizienten der Abrasivität und Mahlbarkeit (Granulats: Essai d’abrasivité et de broyabilité (P 18-579). Frankreich: AFNOR Association francaise de normalisation, Beuth, 2013. [2] DIN 18312, VOB Vergabe und Vertragsordnung für Bauleistungen - Teil C: Allgemeine Technische Vertragsbedingungen für Bauleistungen (ATV) - Untertagebauarbeiten, Beuth-Verlag, 2019. [3] H. Käsling, J. Düllmann and R. J. Plinninger, “Bestimmung der Abrasivität von Festgesteinen mit dem LCPC- Versuch,” Geotechnik 45, no. Heft 2, 2022. [4] Deutscher Ausschuss für unterirdisches Bauen e.V. (DAUB), “Empfehlung zur Auswahl von Tunnelbohrmaschinen,” 2021. [5] M. Feinendegen, T. Babendererde, P. Drucker, J. Holzhäuser, L. Langmaack and A. Richter, “Empfehlung(en) ‘Verschleiß und Verklebung im Lockergestein’ - ein erster Ausblick,” Tagungsband zu Fachsektionstage Geotechnik der DGGT, pp. 268-273, 2023. [6] Z. Sun, Z. Yang, Y. Jiang, H. Gao, K. Fang and M. Yin, “Influence of particle size distribution, test time, and moisture content on sandy stratum LCPC abrasivity test results,” Bulletin of Engineering Geology and the Environment, 17 Juli 2020. [7] M. Abu Bakar, Y. Majeed and M. Rashid, “Influence of propeller material hardness, testing time, rock properties, and conditioning on LCPC rock abrasiveness test,” Bulletin of Engineering Geology and the Environment, 8 Oktober 2020. [8] J. Küpferle, A. Röttger, M. Alber and W. Theisen, “Bewertung des LCPC-Abrasivitätstests aus werkstofftechnischer Sicht,” Geomechanics and Tunneling 8, 2015. [9] K. Thuro, J. Singer, H. Käsling and M. Bauer, “Soil Abrasivity Assessment Using the LCPC Testing Device,” Felsbau 24, 2006. [10] K. Thuro, J. Singer and H. Käsling, “Determining abrasivity with the LCPC Test,” ResearchGate, May 2007. [11] B. Janc, V. Jovicic and Z. Vukelic, “Laboratory Test Methods for Assessing the Abrasivity of Rocks and Soils in Geotechnology and Mining Applications,” Sciendo, 07 September 2020. [12] A. Cheshomi, N. G. Zadeha and A. Sadeghia, “A new approach to LCPC test based on effect of rotation speed (RS) and rotation time (RT),” Arabian Journal of Geoscience, 20 September 2021. [13] M. Hamzaban, J. Rostami, S. Farrokhzadeh, M. Faridazad and F. Rasouli, “A Critique on the Correlations Between Rock Abrasiveness, Hardness, and Mechanical Characteristics: Implication of LCPC Abrasiveness and Breakability Coefficients,” Rock Mechanics and Rock Engineering, November 2024. [14] C. Budach, D. Katrakova-Krüger, B. Siebert and P. Erdmann, “Untersuchungen zur Bestimmung der Abrasivität von grobkörnigen Böden im maschinellen Tunnelbau,” Tagungsband 63. Tribologie-Fachtagung; Reibung, Schmierung und Verschleiß, 26-28 September 2022. [15] C. Budach, K. Lipka, P. Erdmann and D. Katrakova-Krüger, “Ergebnisse von aktuellen Forschungsvorhaben zur Bestimmung der Abrasivität von Lockergestein,” 14. Kolloquium Bauen in Boden und Fels, Januar 2024. [16] C. Budach, N. von Taschitzki, P. Erdmann and D. Katrakova-Krüger, “Prüfung der Abrasivität von Lockergestein: Ergebnisse des Forschungsprojekts ‘Verschleißreduzierung an Werkzeugen von mobilen Arbeitsmaschinen’ (VerA) und des LCPC-Ringversuchs,” Baugrundtagung, 2024. [17] L. Brungs, “Untersuchungen zur Abrasivität grobkörniger Böden im Kontext der aktuellen Normung und Empfehlungen,” TH Köln, Köln, 2023 - unveröffentlicht. [18] J. Arendt, “Erweiterte Untersuchungen zur Abrasivität grobkörniger Böden,” TH Köln, Köln, 2023 - unveröffentlicht. [19] TH Köln, Bericht zum Projekt “Vergleichende Untersuchungen zur Auswahl von Methoden zur Bestimmung der Abrasivität bzw. des Verschleißes bei Abbauwerkzeugen - Teil A”, unveröffentlicht, gefördert durch den Transferfonds der TH Köln, 2022. [20] P. Erdmann, C. Budach, D. Katrakova-Krüger and B. Siebert, “Untersuchungen zur Bestimmung der Abrasivität von grobkörnigen Böden im maschinellen Erd- und Tunnelbau,” 9. Fachtagung Baumaschinentechnik, 2022. [21] D. Katrakova-Krüger, J. Kotscha, C. Budach and P. Erdmann, “Werkstoffuntersuchungen an Prüfflügeln aus dem LCPC-Versuch zur Bestimmung der Abrasivität von Böden,” Tagungsband der 65. Tribologie Fachtagung, ISBN: 978-3-9817451-9-1, 45/ 1-11, September 2024. Science and Research 55 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 DOI 10.24053/ TuS-2024-0038 News 56 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 BOOK RECOMMENDATION expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. 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Content Ignition and inflammation of conventional and alternative fuels such as hydrogen, ammonia, methanol etc. - Combustion processes for alternative fuels - Prevention of irregular combustion phenomena when using conventional and alternative fuels - Methods for measurement and analysis of irregular combustion phenomena - Modern virtual development methods - Control, regulation and latest function algorithms Marc Sens (ED.) 6th International Conference on Ignition Systems for SI Engines 7th International Conference on Knocking in SI Engines 1st edition 2024, 386 p. €[D] 189,00 ISBN 978-3-381-12991-1 eISBN 978-3-381-12992-8 News 57 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 Gerhard Poll was born on February 5, 1954 as the only child of Maria-Anne and Walter Poll in Schweinfurt and if he had dedicated himself to football, he would probably call himself a real “Schnüdel”, as we Lower Franconia know. The High German translation for this is: Zipfel The family has often changed their place of residence due to the professional activity of his father, who was an engineer at SKF. So, in 1956 the family moved from Schweinfurt to Cologne, where Gerhard Poll attended primary school from 1960 to 1964. In 1964 he first transferred to the grammar school in Aachen and - due to another move in 1970 - later to the Celtis Gynasium in Schweinfurt, where he graduated in 1972. So, he only attended school for 12 years, as he was a victim of the so-called short school years. In 1972, he began his mechanical engineering studies at RWTH Aachen University. During the lecture-free period, he worked as an intern at SKF in Schweinfurt as well as at the Talbot wagon factory in Aachen. After completing his intermediate diploma in 1974, Gerhard Poll concentrated more on vehicle engineering, especially for trains. He completed his studies with the diploma thesis entitled “Force-closure investigations on the large rolling test bench” in 1977 under Prof. Krause. He carried out the experimental investigations for his diploma thesis at Thyssen Henschel in Kassel on the rolling test bench (later Bombardier, now Alstom), from which the final research topic of his doctorate was derived. After graduating, he started as a research assistant in the field of “Wear of Materials” under Prof. Hans Krause, who opened the door to the wide world of tribology for him. Gerhard Poll formulated an application within the framework of the Cycle Rail Research Program of the Federal Ministry of Research and Technology. The approval was based on deviations between theoretical predictions and experimental measurement results on test benches and in route tests with real rail vehicles. In 1983, he finally received his doctorate from RWTH Aachen with his dissertation “The influence of real system properties on the force closure laws in rolling relative motion”. Tests were carried out in Munich Freimann on the rolling test bench and on the rolling test bench in Kassel. Among other things, he also used a 2-disc test bench that he had modernized and expanded, which his father had played a major role in designing at the institute in Aachen in the 60s. In 1984, he finally started at SKF in Schweinfurt in technical consulting. Here he first devoted himself to the topics of rail vehicles, shipbuilding and aircraft design, then machine tools, conveyor technology, wind energy, heavy machinery and finally electric machines. During this time, among other things, the commissioning of the 1st ICE and the test dismantling after the first trial period took place. A secondary activity was the recommissioning of steam locomotives for the 150th anniversary of the Deutsche Bahn in 1985 In 1987, he joined SKF in the Netherlands, where he worked at the research center in the field of sealing and tribology. The field of work dealt with spherical roller bearings, wheel bearing units, pillow block housings and seals, mostly in connection with grease lubrication and the application of water and dirt. During this time, he also had close contacts with TU Eindhofen (Harry van Leeuwen), University Delft (Joost Kalker) and University of Twente (Kees Venner) as well as Imperial College (Alistair Cameron). His specialties were the optical lubricating film investigations with by means of fluorescence in sealing contacts and the mechanisms of the backflow effect of seals. This was followed by international lectures, e.g. at STLE in Toronto. In 1992, after SKF acquired seal manufacturer Chicago Rawhide, he moved to Elgin, Illinois in the USA as Head of Research and Product Development. Here, Prof. Poll had worked on seals for truck hub units, for seals for water pumps of diesel engines, and PTFE seals for drying cylinders of paper machines. Decisive and formative during this time was a colleague and getting to know her best friend Lu, who is his wife today. Award of the Vogelphohl Medal of Honour to Prof. Dr.-Ing. Gerhard Poll Laudatio von Cornelia Haag News 58 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 In 1996, Gerhard Poll finally became Professor of Machine Elements and Design Technology and Head of the Institute of Machine Design and Tribology at Leibniz University in Hanover, succeeding Prof. Paland. Under his leadership, research was carried out on rolling bearings and other tribological topics, mostly within the framework of FVA and DGMK research projects. In addition, joint projects funded by the Ministry of Economic Affairs and DFG projects are also organized. The beginnings included projects of the FVA and DGMK such as grease lubrication at low temperatures, grease lubrication of high-speed angular contact ball bearings, friction calculation, and rheological fluid models. Furthermore, work was carried out on rolling bearings for combustion engines and for wind turbines, WEC and electrical load with current passage. The compatibility of elastomers and lubricants was also an ongoing topic. Since 2024, he has been officially retired, but he still advises his temporary successor Prof. Max Marian and gives lectures. In his career, Prof. Poll has published a number of publications, such as the author or co-author of several books, e.g. the book “Mechanics of Solid State Friction”. While still a student, Prof. Krause, his supervisor at RWTH Aachen, left him the documents on the subject of “friction” handed over by Vogelpohl’s widow and the book on friction mechanics was created from them. He has also contributed to the books “Dubbel - Taschenbuch für den Maschinenbau”, “Konstruktionselemente des Maschinenbaus II” and the corresponding exercises. In addition, he is the author or co-author of approx. 350 publications, the last one - for the time being - published in September 2024. He is also the holder of several patents. Prof. Poll has been a member of the GfT since 1997. He was actively involved in the Technical-Scientific Advisory Board, which he had been a member of since 2002 and chaired in 2007. He then moved to the board in 2016, of which he has been a member to date. He is also the chairman of the program committee of the GfT Conference Prof. Poll has always been passionate about tribology and was also very active outside the GfT and founded the conference “Bearing World” as well as the “Bearing World Journal” in 2016 - supported by Prof. Bernd Sauer - and is on the board of the conference and on the editorial board of the journal. He is also a member of the Editorial Board of the Journal of Engineering Tribology. Gerhard’s great role model and promoter is probably his father Walter, who worked as a mechanical engineer in application technology and to whom Gerhard explained the term “tribology” as early as the 60s, which was introduced by Peter Jost at that time. He immediately understood that he had always been a tribologist and explained that he had known this term for a long time, namely from the book “Simplicissimus” by Grimmelshausen. There, the young Simplicissimus was “tribulated” - he got a rubdown. In Bavaria, the term “he gets one bribed” is also used. Through his father, he was born with the hobby of locomotives, water wheels, trams and machines and through him he got to know tribologists and later Vogelpohl honorary award winners at an early age, which has not changed in the course of his further career. You can see the water wheels here at a young age, proudly with the three-phase lock. He also likes to travel by train and he even goes on vacation just to travel by train, e.g. to Switzerland and the family is then taken along without further ado. He also knows individual railway tracks throughout Germany, which he feels like he has already traveled all of them himself. Some characteristic qualities of Prof. Poll should be mentioned here. He is an energetic, creative and enthusiastic researcher and teacher who has achieved great things with a small group. In addition, he also knows how to enlighten his employees in technical discussions with the following words: “Well, there has always been this problem with the railways ...” “and at that time at the railway we have...”. Prof Poll is a good organizer of conferences, symposia and technical discussions and has built up a large network over the years. He is known in the world of tribology both nationally and internationally like a colorful dog and he knew everyone by name. If you want to get in touch, just contact Prof Poll. Furthermore, he is very sociable as you can see and he always has a clear view And if he doesn’t have it, then he knows who to call. He is a caretaker and leaves no one out in the rain, even when things get stormy. He has always stood up for his employees so that they have a job and keep it. This is probably not always to be taken for granted. News 59 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 On the other hand, the doctoral students did not have it easy with him either, before presentations at conferences, the already completed presentations were often discussed again AT NIGHT - in a bathrobe, so to speak - and content was significantly supplemented and expanded. But all in all, Prof Poll is a warm and lovable person. Therefore, he is married and has an 18-year-old son. Prof Poll has, so to speak, a very personal relationship with the namesake of today’s honour - Georg Vogelpohl. In 1944, he passed his habilitation examination at the Technical University of Hanover, where Prof. Poll was the director of the institute for many years. In addition, his widow left him many important notes and many of Prof. Poll’s companions are already bearers of the Vogelpohl Medal of Honour. Prof. Poll has dedicated himself to tribology throughout his career. For his commitment to the field of tribology in industry, research, further education and association work, Prof. Poll was awarded the Georg Vogelpohl Medal of Honour of the German Society for Tribology on 23 September 2024. News 60 Tribologie + Schmierungstechnik · volume 71 · issue 5-6/ 2024 UVK Verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 \ 72070 Tübingen \ Germany \ Tel. +49 (0)7071 97 97 0 \ info@narr.de \ www.narr.de BOOK RECOMMENDATION For years, advocates of professional project work have stressed the growing shift towards project-oriented work structures. This has now become a reality in the daily routines of many employees and managers. Consequently, strong project management skills are becoming increasingly vital to business success. Following the five project management phases of DIN 69901: 2009 and supplemented by chapters on cross-phase competencies and agile methods, this book offers a clear and professionally sound presentation of the modernised ICB 4.0 framework (effective from January 1, 2024). Its structured content, illustrated by a consistent project example, not only guides readers but also ensures they are well-prepared to meet the IPMA ICB 4.0 examination requirements. This book equips anyone seeking to engage in professional project management with the knowledge and tools needed to successfully apply current best practices. Marcus Schulz, Tinka Meier Project Management A Practical Guideline for Today’s Project Managers 3 rd edition 2025, 258 p. €[D] 44,99 ISBN 978-3-381-13481-6 eISBN 978-3-381-13482-3 Checklist Author information Corresponding author: F Mailing address F Telephone and fax number F eMail All authors: F Academic titles F Full name F ORCID (optional) F Research instititute / company F Location and zip code Length F Approximately: 3,500 words Data F Word and pdf documents (both with images + captions) F Additionally, please send images as tif or jpg / 300 dpi / Please send vector data as eps Manuscript F Short and concise title F Keywords: 6-8 terms F Abstract (100 words) F Numbered pictures/ diagrams/ tables (please refer to the numbers in your text) F List of works cited After the typesetting is completed, you will receive the proofs, which you are requested to review and then give your approval to start the printing process. Changes to the manuscript are no longer possible at this stage. 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You can obtain the full open access service for a one-off article processing charge of € 1,850.00 (plus VAT). Editor in chief Dr. Manfred Jungk eMail: jungk@verlag.expert Publisher expert verlag Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 D-72070 Tübingen Tel.: +49 (0)7071 97 556 0 eMail: info@verlag.expert www.expertverlag.de Editor Patrick Sorg eMail: sorg@verlag.expert Tel.: +49 (0)7071 97 556 57 Tribologie und Schmierungstechnik Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie | An Official Journal of Österreichische Tribologische Gesellschaft | An Official Journal of Swiss Tribology We’re looking forward to your contribution! ISSN 0724-3472 Science and Research www.expertverlag.de Timo Hacker, Arne Bischofberger, Katharina Bause, Sascha Ott, Albert Albers Optimization of the run-in and wear behavior of dry-running friction systems through the targeted adaptation of the counter-friction disk: Experimental investigations and potentials for sustainable and efficient solutions Thomas Decker, Georg Jacobs, Julian Röder, Timm Jakobs, Jan Euler Derivation of running-in procedures for planetary journal bearings in wind turbine gearboxes by means of abrasive wear simulation Johannes Wirkner, Mirjam Baese, Astrid Lebel, Charlotte Besser, Hermann Pflaum, Katharina Voelkel, Thomas Schneider, Karsten Stahl Impact of Water Contamination and Iron Particles on the Performance Loss of e-Drive Transmission Fluids in Wet Clutches Thomas Decker, Georg Jacobs, Carsten Graeske, Pascal Bußkamp, Julian Röder, Tim Schröder Model uncertainty of a multiscale, elastohydrodynamic simulation method for the prediction of abrasive wear in journal bearings Laura Stubbe, Yvo Stiemcke, Sarah Mross, Sarah Staub, Konrad Steiner, Kerstin Münnemann, Oliver Koch, Stefan Thielen A new rubber-lubricant compatibility test on a tribometer for radial shaft seals Danka Katrakova-Krüger, Jonas Kotscha, Christoph Budach, Peter Erdmann Materials Investigation of Test Wings from LCPC Abrasiveness Test with Different Soils and Varying Test Duration