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Tribologie und Schmierungstechnik EDITOR IN CHIEF MANFRED JUNGK 3-4 _ 25 VOLUME 72 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 3-4 | 2025 Volume 72 Editor in chief: Dr. Manfred Jungk Tel.: +49 (0)177 1902330 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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Design and layout: Ludwig-Kirn Layout, 71638 Ludwigsburg expert verlag Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5, 72070 Tübingen Tel. +49 (0)7071 97 556 0 eMail: info@verlag.expert Kreissparkasse Tübingen IBAN DE57 6415 0020 0004 7840 30 | BIC SOLADES1TUB USt.-IdNr. DE 234182960 Adverts: eMail: anzeigen@narr.de Tel.: +49 (0)7071 97 97 10 We will gladly send you information and media data. Subscription service: eMail: abo-service@narr.de Tel.: +49 (0)89 85 853 881 Subscription rates: www.meta.narr.de/ zeitschriften/ journals_preisliste.pdf By providing proof of their membership, members of the GfT receive a discount of 20%. Subscription is included for members of the ÖTG. Payment due annually in advance without deduction after the invoice is issued by the publisher. Written cancellation of the subscription is possible until six weeks before the end of the reference year at the latest. Receiving the journal for a reduced price obligates the subscriber to purchase the whole volume. If the subscription is terminated prematurely, the unit price will be charged. Higher power cancels delivery obligation. Place of performance and jurisdiction: Tübingen. ISSN 0724-3472 ISBN 978-3-381-13791-6 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 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0012 As a teenager some connoisseurs told me that the 4 th state of matter for water is tomatoes from green houses in Holland. During Physics classes I learned that Plasma is the 4 th state of matter and commonly known applications are plasma television or plasma etching. The Lawrence Livermore National Laboratory announced at the end of 2022 that after laser pulsing a Deuterium and Tritium filled pellet with 2.05 Megajoule fusion energy of 3.14 Megajoule was released. The fusion releases Helium, Neutrons and Alpha Particles. The latter are absorbed by the Plasma and no radioactivity was formed. Over the last 3 years a lot of funds went into fusion energy research. This summer a 4-year research project on recycling glass fiber reinforced plastics (GRP) utilizing plasma was started. GRP places high demands on the recycling industry due to their complex composition. Currently, GRP waste is predominantly landfilled, incinerated for energy recovery, or used as filler in the cement industry. Special recycling methods are required to ensure that materials such as GRP, which are generated in large quantities from discarded wind turbine rotor blades, can be returned to a circular economy. One of the most promising methods for utilizing difficult-to-recycle composite materials is plasma gasification. The Project Objectives according to the participating partners are to investigate the feasibility and economic viability of plasma gasification for GRP. To this end, the process is being specifically adapted to the specific requirements of GRP, and its technical feasibility is being demonstrated on a laboratory and pilot scale. In thermal plasma, the polymers contained are converted into synthesis gas, while the glass fibers are separated in the form of slag. Heating the water from outside is intended to create an emission-free and residue-free recycling process. The study is investigating whether the secondary raw materials produced are of sufficiently high quality to be used as starting material for new chemical products in the spirit of a cradle-to-cradle cycle. The Leibniz Institute for Plasma Research and Technology (INP), Department of Radiation Technology is tasked for Plasma Gasification on Laboratory Scale and Clarification of Process Mechanisms. The Technical University of Freiberg (TUBAF), Department of Plasma Assisted Conversion handles the Gasification Process in Pilot Plant and Characterization of Product Quality. The Institute for Environment & Energy, Technology and Analytics (IUTA), Department of Resources & Recycling Technology contributes mechanical pretreatment and specification along the entire process chain. 1 million tons of GRP were produced 2022 in the European Union. Regardless if the consumed energy for this new process can be offset or the Wind Turbine share of GRP is high the public acceptance of Wind Turbines will increase. The nimby (not in my back yard) attitude slowed down the growth of wind turbines and fake news like wind turbines after use create toxic waste have influenced the public acceptance. If that funding went into tribology would be an alternative, leaves me to remind you Tribology is everywhere. Your editor in chief Manfred Jungk Plasma-Technology for Wind Turbines Events 2 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 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 ► 26.01. - 30.01.26 Khajuraho, India INF Nanotribolog ► 27.01. - 29.01.26 Ostfildern, Germany 25 th International Colloquium Tribology ► 26.03. - 27.03.26 Bangkok, Thailand Fuels and Lubes Asia Week ► 11.04. - 14.04.26 Rome, Italy ELGI Annual General Meeting ► 21.04. - 22.04.26 Stuttgart, Germany UNITI Mineral Oil Technology Congress ► 17.05. - 21.05.25 Orleans, Louisiana (USA) 80 th STLE Annual Meeting & Exhibition Contents 3 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 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 72, Issue 3-4 December 2025 35 Wilhelm Rehbein, Isabell Lange, Kevin DiNicola, Salvatore Rea, John Williams EP Additives with Enhanced Sustainability for Water-miscible Metalworking Fluids 42 Markus Grebe, Henrik Buse, Richard Heinlein, Andreas Keller Modern Application-Oriented Tribometry: Understanding Tribology Instead of Producing Characteristic Values - How Modern Measurement Technology Can Enable a View into the Hidden Tribological Contact 50 Katrin Alt, Markus Wöppermann, Frank Plenert, Jürgen Liebrecht Design and Implementation of a Model Test for Reproducing Thickener Degradation in Grease-Lubricated Gearboxes and Its Role in Grease Formulation Development 54 Ricardo Crespo Martins, Dennis Konopka, Mareike Dukat, Florian Pape, Martin Nicolaus, Kai Möhwald, Gerhard Poll, Max Marian Molybdenum APS-Coatings: A Selfregenerative Solution for Wear Resistance in Dry-lubrication Applications 59 Hoang Viet Le, Kai Weigel Influence of tool coatings on the tribological effectiveness of bio-lubricants for sheet metal forming 64 Adrian Heinl, Erich Prem, Christian Wilbs, Fabian Kaiser, Daniel Frölich Optimization of Radial-Shaft-Sealing systems for the use at high circumferential speeds 71 Dennis Jess, Andreas Ligocki Concept for the generation of single-type wear particles for training the AI-based image processing of a particle sensor 82 News Österreichische Tribologische Gesellschaft Columns 5 Merle Reimers, Silvia Richter, Georg Jacobs, Joachim Mayer, Florian König Influence of Corrosion Inhibitors on the Wear Protection of Extreme Pressure / Anti Wear Additives in Oil-lubricated Rolling Bearings 14 Nadja Aufderstroth, Lennart Schierholz, Jaacob Vorgerd, Manuel Oehler Test method using non-circular discs with a locally varying slide-to-roll ratio to investigate scuffing 22 Christian Spura Wear energy density for wear prediction of displaceable spline couplings 31 Felix Bernhardt, Katrin Alt, Markus Wöppermann Systematic investigation of the µ-mechanical material change of the sealing edge of radial shaft seals 1 Editorial Plasma-Technology for Wind Turbines 2 Events Science and Research 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 72 · issue 3-4/ 2025 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 This monograph takes a new look at tribology with its basic concepts of friction and wear using the example of lubricating greases. The consideration of the phenomenon of occurring instabilities and the introduction of the entropy concept into lubricating grease tribology provide a new perspective on known phenomena. The second part of this book presents a wide range of experimental possibilities for investigating lubricating greases. Contents Introduction to Instability and Postmodern Tribology - On the Phenomenon of Self - Organization - Postmodern Grease Tribology - Lubricating Grease - Rheological behavior of Lubricating greases - A Selected Traditional Wear Model - The Extension of the Wear Concept Erik Kuhn On the Tribology of Lubricating Greases An energetic approach to post-modern tribology Tribologie - Schmierung, Reibung, Verschleiß 1st edition 2025, approx. 210 p. €[D] 118,00 ISBN 978-3-381-14171-5 eISBN 978-3-381-14172-2 1 Introduction In rolling element bearings, lubricants such as oils and greases are used to reduce friction and wear. In addition, these lubricants also fulfil other tasks, such as preventing corrosion on the metal surfaces of the bearings. As these requirements cannot be adequately fulfilled by base oils alone, additives are added, which specifically improve the lubricant properties. In order to achieve the desired properties, various additives are used simultaneously in one lubricant. The simultaneous use of various surfaceactive additives might cause interactions between those additives, while forming a tribological boundary layer. Synergistic as well as antagonistic effects of interactions can occur that potentially influence the wear protection performance of the additives. For example, the performance of zinc dialkyl dithiophosphate (ZDDP) can be enhanced by friction modifier additives [1,2]. In contrast, by adding other additives such as dispersants the tribological boundary layer formation rate of ZDDP can be reduced [3,4]. The reduction or at least knowledge of harmful additive interactions is required to ensure optimum use of the additives' potential. Likewise, an increase in the positive additive interactions can be used to reduce the needed quantity of additives. Therefore, the mechanisms of the surfaceactive additives need to be investigated in combination with each other. Previous studies with the sole use of the extreme pressure / anti wear (EP/ AW) additive ZDDP revealed that a homogeneous layer on the rolling elements is desired to assure good wear protection [5]. However, commercial oils commonly contain extreme pressure / anti wear (EP/ AW) and corrosion inhibitor (CI) additives, which led to antagonistic interactions [6,7]. Based on the mechanisms Science and Research 5 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0013 Influence of Corrosion Inhibitors on the Wear Protection of Extreme Pressure / Anti Wear Additives in Oil-lubricated Rolling Bearings Merle Reimers, Silvia Richter, Georg Jacobs, Joachim Mayer, Florian König* submitted: 19.09.2024 accepted: 08.08.2025 (peer review) Presented at GfT Conference 2024 Surface-active additives have a central influence on the wear behavior of rolling bearings. The simultaneous use of different surface-active additives can cause interactions between these additives and influence the formation of a tribological boundary layer required for wear protection. Until now, these interactions between corrosion inhibitors (CI) and extreme pressure / anti wear additives (EP/ AW) have been insufficiently investigated. This article therefore presents a method for evaluating the influence of CI on the effect of EP/ AW additives in oil-lubricated rolling bearings. In standard FE8 tests according to DIN 51819-3, the mineral oil used with the EP/ AW additive zinc dialkyl dithiophosphate (ZDDP) shows insufficient wear protection. By adding the corrosion inhibitor zinc carboxylate, the wear mass can be significantly reduced but not completely eliminated under the same conditions. To evaluate the formation of the tribological boundary layers, further tests are therefore carried out at a higher relative lubricant film height. The influence on the composition of the boundary layer is analyzed using electron probe microanalysis (EPMA). By adding zinc carboxylate, a higher mass coverage can be achieved in the tribological boundary layer than when using ZDDP alone. Keywords Wear, Additives, Lubricant, Additive Interactions, Extreme Pressure / Anti Wear Additive, Corrosions Inhibitor, Rolling Bearings Abstract * Merle Reimers a (corresponding author) Orcid-ID: https: / / orcid.org/ 0000-0002-8880-5561 Dr. rer. nat. Silvia Richter b , Dr.-Ing. Georg Jacobs a , Dr. rer. nat. Joachim Mayer b , Dr.-Ing. Florian König a a Institute for Machine Elements and Systems Engineering RWTH Aachen University Schinkelstr. 10, 52062 Aachen, Germany b Central Facility for Electron Microscopy RWTH Aachen University Ahornstr. 55, 52074 Aachen, Germany neral oil and 1.0 wt.% ZDDP is defined as the reference. Therefore, three oils were investigated in this study, which are shown in Table 1. The ZDDP used in this study is a primary ZDDP with a C8 chain. It contains 8 wt.% phosphorus, 16 wt.% sulfur and 9.5 wt.% zinc. The calcium sulfonate is composed of 2 wt.% calcium and the zinc carboxylate holds 15 wt.% zinc. 2.2 FE8 Test Method To evaluate the influences of CI on the wear protection of EP/ AW under high loads axial thrust bearing tests (FE8 tests) based on DIN 518193 [8] are carried out. A FE8 test rig with axial cylindrical roller bearings (81212) was used for this purpose. The bearings each consist of 15 rolling elements, two washers and a cage. All rolling elements and washers are made of 100Cr6, while a polyamide cage (PA66) was selected in this study to avoid the chemical effects associated with other cage materials such as brass. Additionally, only rolling elements and washers with an arithmetic mean roughness value Ra = 0.04 - 0.05 µm were used, to eliminate potential influences of varying roughness on the wear test results. Two of the test bearings described are tested simultaneously and loaded by a hydraulic load and a disc spring assembly. Each bearing is individually supplied with 0.1 liters of oil per minute. The shaft with the two bearings is driven at the desired rotational speed with continuous measurement of the frictional torque. A heating ja- Science and Research 6 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0013 of EP/ AW, it is assumed that the joint use of EP/ AW and CI hinders the forming of homogeneous, tribological boundary layers. Consequently, to ensure a safe operation while using both EP/ AW and CI the interactions need to be better understood. The aim of this contribution is to introduce a method to evaluate the influences of the CI on the wear protection of EP/ AW when simultaneously using EP/ AW and CI in oillubricated rolling bearings. Therefore, axial thrust bearing tests (FE8 tests) accordingly to DIN 518193 [8] are carried out to provide a practical assessment for highly loaded rolling contacts. The base oil with only the EP/ AW is set as reference. For the simultaneous use, the EP/ AW is combined with either one of the two CI of this study. The influences of CI on the wear protection of EP/ AW are evaluated by analyzing the wear mass of the rolling elements sets. Furthermore, electron microscopic methods such as electron probe microanalysis (EPMA) are used to investigate the chemical compositions of tribological boundary layers. This allows the influence of the CI on the tribological boundary layers to be examined. Specifically, the mass thickness and chemical composition of tribological boundary layers formed in the FE8 tests are measured. 2 Materials and Methods The oils used, including the base oil and additives, are presented below. The test method and microanalysis methods are also described. 2.1 Base Oil and Additives The oils consist of mineral oil (API group I) as base oil, 1.0 wt.% zinc dialkyl dithiophosphate (ZDDP) as EP/ AW and either 0.5 wt.% calcium sulfonate or 0.5 wt.% zinc carboxylate as CI. The oil with only mibase oil 1.0 wt.% EP/ AW 0.5 wt.% CI M-ZDDP mineral oil (API group I) ZDDP - M1 calcium sulfonate M2 zinc carboxylate Table 1: Oils used in this study a) Axial thrust bearing test rig (FE8 test rig) oil supply axial load heating and cooling drive 2 test bearings torque measurement 0 -5 +5 0 +5 -5 0 -12.05 +13.70 length [mm] SRR [%] b) Test bearing Figure 1: FE8 test rig and test bearing cket is used to maintain the required test temperature of the oil. The test rig and the bearing type are shown in Figure 1 a. Due to the different linear speeds at the inner and outer end of the cylindrical rolling elements, different slide-roll-ratios (SRR) are present along its length. This leads to different products of relative velocity and pressure along the rolling element, which affects the formation of the tribological layers [9]. This is therefore also included in this study and shown in Figure 1 b. All FE8 tests were conducted under the standardized conditions of 80 kN (1.89 GPa), 80 °C and 7.5 rpm for 80 hours. Each oil was tested once due to the good reproducibility as shown in [5,10]. Furthermore, all FE8 tests were conducted with two test bearings and each bearing is analyzed separately. The FE8 test is analyzed on the basis of the wear mass of rolling element sets consisting 15 rolling elements. The wear mass is determined by weighing the rolling element sets of each test bearing before and after the test with an accuracy of 1~mg. The mean value from the rolling element sets of both test bearings is compared in this work in order to evaluate the influences of CI on the wear protection of EP/ AW. In addition, the total wear masses of the rolling element sets per bearing are given as error bars. Based on [11] a wear mass less than 10 mg is rated as very good wear protection of the oil, when tested under 80 kN, 80 °C and 7.5 rpm for 80 hours. A test result with a wear mass less between 10 mg and 30 mg is still considered good, while a wear mass over 100 mg is rated as very high wear, i.e. very poor wear protection of the oil. 2.3 Microanalysis Methods After the FE8 tests the rolling elements were subjected to a cleaning procedure to remove residues of the lubricant film. The cleaning procedure was established by checking the reproducibility of the thin film analysis. Thus, the cleaning recipe can be described as follows: • Clean all sides of using absorbent cotton and white spirit • Immerse in white spirit for approx. 10 seconds • Clean again with absorbent cotton and white spirit • Dry in the air • Clean with ethanol and acetone in an ultrasonic bath ◦ 5 min ethanol ◦ 5 min acetone ◦ 5 min ethanol ◦ 5 min acetone ◦ 5 min fresh ethanol • Dry with a hair dryer EPMA measurements were carried out with a JEOL JXA-8530F microprobe analyzer equipped with a Schottky Field Emitter and 5 wavelength dispersive spectrometer. The operating conditions applied throughout the present work were: Primary electron energy: 10 keV; beam current: 100 nA; counting time: 10 s/ measuring point. The intensities of O Kα, Zn Lα, P Kα, S Kα and Ca Kα were detected. A synthetic multilayer for the analysis of O was chosen. For the measurements the electron beam was defocused to 10 µm. Thus, local inhomogeneities in the elemental distribution and effects from surface roughness were compensated. The X-ray intensities were calibrated by measurement on the following standards under constant conditions: Fe 2 O 3 for O Kα, Zinc for Zn Lα, GaP for P Kα, FeS 2 for S Kα and Andradite for Ca Kα. The k-ratios for all elements, i.e. the calibrated net intensities from the sample to those of bulk standards, were measured across the sample surface by conventional step scans with a step size of 140 µm. Science and Research 7 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0013 0 -5 +5 +4 0 Mass coverage [μg/ cm 2 ] -4 0 30 15 exemplary EPMA result of rolling element 0 -9.8 +10.8 length [mm] SRR [%] Oxygen Phosphorus Sulfur Calcium Zinc Total Elements Figure 2: Representation of the EPMA results very high (> 1300 mg). Nevertheless, the wear masses of the oil M1 were in the same range of the oil with ZDDP only (MZDDP). The high wear masses of the oil M-ZZDP align with the results of [13]. In contrast, the wear mass of the rolling elements sets was significantly reduced to 88 mg when ZDDP was used in combination with zinc carboxylate (M2), which means a reduction by more than 90 %. For a deeper understanding of the wear protection depending on the type of oil and used additives, the EPMA results performed along the surface of the rolling elements are shown in Figure 4. There is no homogeneous tribological boundary layer formed in FE8 tests for both oils with ZDDP only (M-ZDDP) as well as with ZDDP and calcium sulfonate (M1). Especially, in the first half of the measurement scan (-5 to 0 mm) of the oil with CI (M1) the partial mass coverages of all elements vary strongly. In general, the concentration of additive elements in the total mass coverage is small. The second half of the scan (0 to +5 mm) reveals almost no layer formation with both oils. Most of the surface area is covered by a total mass of about 1 µg/ cm 2 which corresponds to about 2 nm thickness assuming a density of 5 g/ cm 3 . This is the thickness of a natural oxide layer. If zinc carboxylate is added to ZDDP (M2), the EPMA profiles are more homogeneous. Based on previous results, a homogeneous boundary layer is necessary for a properly working wear protection [5]. This correlates with the results of the FE8 test (Figure 3). There are small variations of the mass coverage depending on the SRR condi- Science and Research 8 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0013 The electron beam energy was adjusted to 10 keV to get a surface-sensitive excitation of the interesting tribolayer. The information depth ranges from about 250 to 450 nm depending on the chosen X-ray line. Since the information depth is larger than the expected thickness of the tribological boundary layer a special algorithm (inhouse script) was applied to determine elementspecific mass coverages from the measured data, i.e. k-ratios. As physical model the work of Pouchou and Pichoir [12] was used to convert experimental k ratios into elementspecific or partial mass coverages. The EPMA measurements were carried out along the rolling element length as shown in Figure 2. 3 Results The wear masses of the rolling element sets from the FE8 tests are used to evaluate the influences of the CI on the wear protection of EP/ AW. Therefore, the results of the oil including EP/ AW and CI must be compared with the oil containing only EP/ AW and no CI. The wear masses in mg for the three different oils are shown in Figure 3. The wear mass of the reference test M-ZDDP, which contains solely EP/ AW ZDDP is shown twice for better comparison. None of the three oils lead to a wear protection under the standard FE8 test conditions (7.5 rpm, 80 kN, 80 °C, 80 h, λ = ~ 0.05), since all wear masses were higher than 30 mg. Especially, the wear masses of the oil containing ZDDP and calcium sulfonate (M1) and the oil with ZDDP only (MZDDP) were EP/ AW EP/ AW + CI T = 80 ˚C n = 7.5 rpm F = 80 kN t = 80 h (1.89 GPa) ZDDP zinc dialkyl dithiophosphate Ca-Sulf. calcium sulfonate Zn-Carb. zinc carboxylate standardized FE8 test conditions Oil name M-ZDDP M1 M-ZDDP M2 EP/ AW ZDDP ZDDP CI - Ca-Sulf. - Zn-Carb. Figure 3: Wear masses of FE8 tests with standardized test conditions Science and Research 9 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0013 T = 80 ˚C n = 7.5 rpm F = 80 kN t = 80 h (1.89 GPa) standardized FE8 test conditions M-ZDDP M1 M2 Oxygen Phosphorus Sulfur Calcium Zinc Total Elements Figure 4: EPMA results of rolling elements after FE8 test with standardized test conditions EP/ AW EP/ AW + CI ZDDP zinc dialkyldithiophosphate Ca-Sulf. calcium sulfonate Zn-Carb. zinc carboxylate T = 80 ˚C n = 75 rpm F = 80 kN t = 80 h (1.89 GPa) adapted FE8 test conditions EP/ AW ZDDP ZDDP CI - Ca-Sulf. - Zn-Carb. Figure 5: Wear masses of FE8 tests with adapted test conditions tribological boundary layer enriched with additive-elements has been formed. Due to the adapted FE8 test conditions the influence of the CI on the formation of the layer can be now investigated. In case of the calcium sulfonate as CI an additional Ca mass coverage is visible (M1) compared to the tribological boundary layer of the oil with ZDDP only (M-ZDDP). By adding zinc carboxylate, a higher mass coverage can be found in the tribological boundary layer (M2) compared to the only use of ZDDP (M-ZDDP). In addition, although there is no P portion in both CI used, the P mass coverage increases as well by adding either one of the CI. The CI might have a catalytic effect on the enrichment of P in the boundary layer. 4 Discussion No wear protection was achieved with none of the three oils used in this study at the standardized FE8 test conditions with 7.5 rpm. However, an influence of the CI on the wear protection performance of the EP/ AW can be observed. By adding calcium sulfonate as CI (M1) the wear mass was the same as with ZDDP only (M-ZDDP). In contrast, by using ZDDP with zinc carboxylate simultaneously (M2) over 90 % less wear mass was detected Science and Research 10 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0013 tions along the surface of the rolling element. It points to the formation of a tribological boundary layer, probably iron oxide layer, since the concentration of the additive elements is low. The thickness varies between 20 and 40 nm assuming a density of 5 g/ cm 2 . To investigate the tribological boundary layer closer, the FE8 test conditions were adapted accordingly to [13]. Except for the rotational speed, which is raised to 75 rpm, the test conditions stayed the same. Even with the adapted FE8 test conditions, testing continues in mixed friction (λ = ~0.2). All three oils were investigated again under the described adapted FE8 test conditions with higher rotational speed. The wear masses are shown in Figure 5. For all three oils the wear masses of the rolling elements sets were much lower compared to the standard FE8 test conditions. This can be explained by the less critical FE8 test conditions. Also, the wear masses do not differ significantly between the oils. A slight increase of the wear mass can be seen by adding calcium sulfonate to the oil with ZDDP (M1). The objective to investigate the tribological boundary layer closer by adapting the FE8 test conditions accordingly to [13] was successful. A homogenous tribological boundary layer was formed and the concentration of all additive elements is also higher for all three oils. A T = 80 ˚C n = 75 rpm F = 80 kN t = 80 h (1.89 GPa) adapted FE8 test conditions M-ZDDP M1 M2 Oxygen Phosphorus Sulfur Calcium Zinc Total Elements Figure 6: EPMA results of rolling elements after FE8 test with adapted test conditions compared to using ZDDP only (M-ZDDP). This correlates with the EPMA measurement. No homogenous tribological boundary layer was formed in FE8 tests for both oils with ZDDP only (M-ZDDP) as well as with ZDDP and calcium sulfonate (M1). In contrast, by adding zinc carboxylate to the oil with ZDDP (M2) the EPMA results show a more homogeneous boundary layer, which is necessary for a properly working wear protection [5]. Since the concentration of the additive elements is low, the tribological boundary layer is probably an iron oxide layer. The more homogeneous boundary layer and the lower wear masses indicates a positive influence of the CI zinc carboxylate on the wear protection of ZDDP. Nevertheless, to study the influence of CI on the wear protection of EP/ AW analyzing the boundary layer is needed. By adapting the FE8 test conditions to higher rotational speed (75 rpm) accordingly to [13] the boundary layer can be analyzed successfully. In comparison to the results of the adapted FE8 test with the oil containing of ZDDP only (MZDDP) the wear mass of the oil also consisting of calcium sulfonate (M1) slightly increases. This indicates a negative influence of the CI calcium sulfonate on the wear protection of ZDDP. The EPMA results of M1 show a calcium coverage in the tribological boundary layer as well as a higher phosphorus coverage compare to the EPMA results of MZDDP. Overall, this study both showed standardized as well as adapted FE8 tests are necessary to analyze the influence of CI on the wear protection of EP/ AW in oillubricated rolling bearings. The FE8 tests with the standardized test conditions (80 kN, 80 °C, 7.5 rpm, 80 hours) are needed to examine the wear protection by means of wear masses. Since no homogeneous tribological boundary layer was formed under the standardized FE8 test conditions, a closer investigation of the influences of the CI on the wear protection of EP/ AW was not possible. Therefore, the adapted test conditions (80 kN, 80 °C, 75 rpm, 80 hours) are used. Homogeneous tribological boundary layer were formed and can be further analyzed with microanalysis methods such as EPMA. Even if the wear masses are lower in the adapted FE8 test due to the less critical operating conditions, a slight influence of the CI on wear protection of EP/ AW can be seen. Nevertheless, the results from the standardized FE8 tests are necessary for a conclusive study. 5 Conclusion While forming a tribological boundary layer, synergistic as well as antagonistic effects of interactions between surfaceactive additives might occur and potentially influence the performance of those additives. The surfaceactive additives extreme pressure / anti wear (EP/ AW) and corrosion inhibitor (CI) additives are assessed as antagonistic [6,7]. However, EP/ AW and CI are used simultaneously in oillubricated rolling bearings. To ensure a safe operation of the rolling bearings, the influence of the CI on the performance which have been insufficiently investigated. A method to evaluate the influences of the CI on the wear protection of EP/ AW when using EP/ AW and CI simultaneously in oillubricated rolling bearings was introduced. Therefore, standard FE8 tests were carried out. Under standard FE8 test conditions, none of the three oils lead to a sufficient wear protection. Nevertheless, the wear mass was significantly reduced by adding zinc carboxylate compared to the sole use of ZDDP. This can be explained by the more homogeneous tribological boundary layer formed by this additive combination. To investigate the tribological boundary layer closer, the FE8 test conditions were adapted by raising the speed accordingly to [13]. Due to the less critical conditions the wear masses reduced significantly and are in close range to each other. Therefore, the FE8 test with the standard test conditions are needed to evaluate the influence of CI on the wear protection of EP/ AW. Nevertheless, the adapted FE8 test conditions are required as well to investigate the influence on the formed tribological boundary layer using electron probe microanalysis. By combining the standard and the adapted FE8 test with accelerated speed, a method was developed to investigate the influence of the CI on the wear protection of EP/ AW when using EP/ AW and CI simultaneously in oillubricated rolling bearings. The method includes an evaluation of the wear masses and EPMA results of both tests. With the introduced method a first step is shown to evaluate the influences of the CI on the performance of EP/ AW when using EP/ AW and CI simultaneously in oillubricated rolling bearings. To study the synergistic and antagonistic effects of the simultaneously use of EP/ AW and CI further investigations such as variation in different additives, base oils, additive concentration and ratio as well as runningin processes are needed. Acknowledgements The project (01IF22309N) is funded by the Federal Ministry of Economic Affairs and Climate Action BMWK (Bundesministerium für Wirtschaft und Klimaschutz) on the basis of a resolution of the German Bundestag. References [1] Eickworth J, Aydin E, Dienwiebel M, Rühle T, Wilke P, Umbach TR. Synergistic effects of antiwear and friction modifier additives. ILT 2020; 72(8): 1019-25. https: / / doi.org/ 10.1108/ ILT-07-2019-0293. [2] Lu R, Shiode S, Tani H, Tagawa N, Koganezawa S. A Study on the Tribofilm Growth and Tribological Properties Science and Research 11 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0013 schmierstoff-Prüfgerät FE8 - Teil 3: Verfahren für Schmieröl - einzusetzende Prüflager: Axialzylinderrollenlager. Berlin: Beuth Verlag GmbH. [9] Stratmann A. Einflüsse auf die tribologische Grenzschichtbildung beim Betrieb von Wälzlagern in der Mischreibung. Dissertation, Institut für Maschinenelemente und Systementwicklung, Verlag Mainz; RWTH Aachen University. [10] Burghardt G, Wächter F, Jacobs G, Hentschke C. Influence of run-in procedures and thermal surface treatment on the anti-wear performance of additive-free lubricant oils in rolling bearings. Wear 2015; 328-329: 309-17. https: / / doi.org/ 10.1016/ j.wear.2015.02.008. [11] van de Sandt N. Gebrauchsdauer von axial belasteten Wälzlagern bei starker Mischreibung. 1. Aufl. Aachen: Mainz 2004. [12] Pouchou JL, Pichoir F. Electron probe X-ray microanalysis applied to thin surface films and stratified specimens. Scanning Microscopy 1993; 1993(7): 167-89. [13] Burghardt G. Wirkung tribologischer Grenzschichten in Wälzlagern unter Mischreibung. Dissertation, Institut für Maschinenelemente und Systementwicklung, RWTH Aachen University 2017. Science and Research 12 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0013 of Tribofilms Formed from Zinc Dialkyl Dithiophosphate (ZDDP) and Molybdenum Dialkyl Dithiocarbamate (MoDTC). Tribology Online 2018; 13(3): 157-65. https: / / doi.org/ 10.2474/ trol.13.157. [3] Papay AG. Antiwear and extreme-pressure additives in lubricants. Lubrication Science 1998; 10(3): 209-24. https: / / doi.org/ 10.1002/ ls.3010100304. [4] Barcroft FT, Park D. Interactions on heated metal surfaces between zinc dialkyldithiophosphates and other lubricating oil additives. Wear 1986; 108(3): 213-34. [5] Rosenkranz L, Richter S, Jacobs G, et al. Influence of temperature on wear performance of greases in rolling bearings. ILT 2021; 73(6): 862-71. https: / / doi.org/ 10.1108/ ILT-03-2021-0076. [6] Mortier RM, Fox MF, Orszulik ST. Chemistry and Technology of Lubricants. Dordrecht: Springer Netherlands 2010. [7] Spikes HA. Additive-additive and additive-surface interactions in lubrication. Lubrication Science 1989; 2(1): 3- 23. https: / / doi.org/ 10.1002/ ls.3010020102. [8] DIN 51819-3: 2016-12, Prüfung von Schmierstoffen - Mechanisch-dynamische Prüfung auf dem Wälzlager- Science and Research 13 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 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! Ost昀ldern/ Stuttgart, Germany 27th - 29th January 2026 contact pressures and sliding speeds under friction, lead to the collapse of the lubricating film. Without lubrication, the contacting surfaces become bonded, effectively welding together. Due to the rolling motion, these weldings are immediately torn apart, resulting in a roughened surface. The failure mode is characterized by scuffing marks that align with the rolling direction. Adhesive effects cause material to be removed from the surface and transferred to the counter body. Over time, roughened areas and transferred material become partially run in after a certain number of load cycles. [CZI20; LIN22; NIE03; SOM18; VOR23] Lubricants are evaluated for their scuffing load-carrying capacity using an FZG gear test machine, following the test procedures specified in DIN ISO 14635-1 [DIN06]. These test methods provide the basis for determining characteristic values of lubricants which are essential for calculating the scuffing load-carrying capacity for gears. The tests employ the standardized FZG test gears type A Science and Research 14 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0014 1 Introduction A variety of standardized test procedures exists for investigating lubricants. In addition to gears, analogy tests are conducted using test specimens with a simpler geometry, such as the tribological subsystems ball-on-ball [DIN15], ball-on-disc [PCS24] or disc-on-disc [OPT24]. On a tribological view, gear meshing is characterized by varying slide-to-roll ratios. In contrast, most of these subsystems only accommodate constant slide-to-roll ratios. The closest approximation to gear meshing is achieved by using discs as specimens but limitations arise due to the stationary contact conditions. This article introduces new test specimens featuring a variable disc radius. These non-circular discs enable the implementation of locally varying slide-to-roll ratios, offering a higher degree of similarity to gear meshing compared to conventional disc contacts. The research includes tests with non-circular discs on a two-disc tribometer and reference tests with gears on an FZG gear test machine, both evaluating the scuffing load-carrying capacity of various test oils. The results of these tests are discussed in this article. 2 State of research and technology Scuffing is a type of failure that occurs spontaneously due to the loss of lubrication from temporary overloading. The wear mechanism responsible for scuffing is adhesion. Excessive contact temperatures, caused by high Test method using non-circular discs with a locally varying slide-to-roll ratio to investigate scuffing Nadja Aufderstroth, Lennart Schierholz, Jaacob Vorgerd, Manuel Oehler* submitted: 20.09.2024 accepted: 11.08.2025 (peer review) Presented at GfT Conference 2024 Lubricant and wear tests are conducted using both test-specific gears and specimens such as balls or discs. To enhance the disc-disc subsystem, a test specimen was developed whose geometry allows for locally varying slide-to-roll ratios similar to those in gear meshing while maintaining a constant shaft speed. Scuffing load-carrying capacity tests were performed with non-circular discs on a two-disc tribometer and with gears on an FZG gear test machine, following the FZG test procedures. The results demonstrated a strong correlation between the disc and gear tests. In the process, the scuffing load-carrying capacity classes for various lubricants were determined. Additionally, the pv values at the damage points were equivalent for both setups. Keywords discs, analogy tests, two-disc tribometer, slide-to-roll ratio, scuffing load-carrying capacity, adhesion Abstract * Nadja Aufderstroth, M.Sc. (corresponding author) ORCID-ID: https: / / orcid.org/ 0009-0001-9160-3442 Lennart Schierholz, M.Sc. ORCID-ID: https: / / orcid.org/ 0009-0003-5581-233X Dr.-Ing. Jaacob Vorgerd ORCID-ID: https: / / orcid.org/ 0000-0003-0232-2479 Prof. Dr.-Ing. Manuel Oehler ORCID-ID: https: / / orcid.org/ 0000-0001-8251-0896 Lehrstuhl für Antriebstechnik (ante) Fakultät für Maschinenbau, Ruhr-Universität Bochum Universitätsstr. 150, 44801 Bochum which are noted for their uneven slide-to-roll ratio. At the contact point D, the maximum contact pressure and a slide-to-roll ratio of 106 % are observed (Eq. 1). (1) In the development of lubricants, numerous tests are conducted using analog test specimens which enable a cost-efficient and flexible testing. Established tribometers primarily employ the disc-on-disc [OPT24] or ballon-disc [ING15; LI13; PCS24] sub-systems to simulate the tribological conditions. Analogy tests have proven valuable for determining the scuffing load-carrying capacity of lubricants [CON23], the influence of surface treatments [ALN04; PAT95], and the frictional properties [CIH23; GRE22]. However, their transferability is limited due to the stationary contact conditions. During these tests, moderate slide-to-roll ratios (SRR < 50 %) are typically used under constant load conditions. Particularly at high slide-to-roll ratios, substantial cooling is necessary to stabilize the thermal conditions due to heat emission [SAV17]. Analogy tests using the disc-on-disc subsystem typically operate with stationary slide-to-roll ratios. To better replicate gear meshing, it is desirable to achieve locally varying slide-to-roll ratios. One approach involves using a tribometer with elliptic gears to create variable slide-toroll ratios in the contact between two discs with constant radii, which has been used to investigate the scuffing load-carrying capacity [BRE17]. Another approach is to adapt the disc geometry to achieve variable slide-to-roll ratios, for which a corresponding model test rig has been developed [TEN16]. Micro pitting tests were successfully carried out on this test rig [SCH24; TEN22]. 3 Modelling of non-circular discs In gear meshing, locally varying speeds influence the tribological stress and can lead to gear damage such as = 2 ∙ s Σ = 2 ∙ ( t1 − t2 ) t1 + t2 scuffing [CZI20; LIN22]. The disc geometry is designed based on the law of gearing, allowing for the modelling of specimens to generate locally varying slide-toroll ratios during one full rotation. Figure 1 illustrates the model of the non-circular discs. The distance between the centers of rotation O 1 and O 2 corresponds to the center distance a. The pressure angle α defines the angle between the center distance and the contact normal. The pitch point C is located in the center between O 1 and O 2 . The base circle radii r b1 and r b2 are perpendicular to the contact normal. The contact point K, whose contact plane is orthogonal to the contact normal, moves along this line. The minimal radii r min1 and r min2 serve as the starting values for the disc contour generation and are located at the disc angle φ = 180°. The disc contour is symmetric about the x-axis and extends from φ = 0° to φ = 180° for each side. In this article, the basic concept is presented, while the derivation of the disc geometry and its modelling are described in detail in [AUF25]. To ensure a continuous and smooth meshing, the principles of cam design are applied in contour calculations (Figure 1). The infinitesimal change of the contact point must occur within the contact plane (Eq. 2). [ANG91; ROT04]. (2) The contact radius r 1 is calculated using the base circle radius r b1 and the angles α and δ 1 (Eq. 3). The base circle radius r b1 is determined based on the center distance a and the pressure angle α (Eq. 4). (3) (4) The angle δ 1 represents the relative angle from the contact normal to the contact radius r 1 (Eq. 5). The contact 1 1 = 1 ∙ tan ( + 1 ) 1 = b1 sin ( + 1 ) b1 = 0,5 ∙ ∙ sin ( ) Science and Research 15 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0014 Figure 1: Geometric model of the non-circular discs with a constant contact normal and depiction of the contact point motion within the contact plane (6) (7) To adjust the relative speed ratios, the non-circular discs are scaled according to the size ratio u (Eq. 8). (8) The pressure angle over a full disc rotation is shown in Figure 2. Between φ A and φ E , the pressure angle remains constant with a value of 22.5°, and is equal to the normal pressure angle α n . In this region, the contact point moves along the contact normal, and the disc contour follows an involute shape. To ensure a continuous rotary motion of the disc, the disc contour from 0° to 180° is mirrored. In order to connect these regions, a transition function is required that fades smoothly. For this purpose, a continuously differentiable Bézier function is generated based on interpolation points, where the final interpolation point determines the target pressure angle which corresponds with α n . For the scuffing tests, the disc variant S70-A is used, for which the specifications are shown in Table 1. In this 1 , 2 = + 1 , 2 2 = � ( − 1 ∙ cos ( 1 )) 2 + ( 1 ∙ sin ( 1 )) 2 = min1 min2 Science and Research 16 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0014 radius r 2 is then determined using δ 1 (Eq. 6). The angle ξ represents the deviation from the x-axis to the contact radii and is derived from the sum of the angles φ and δ (Eq. 7). (5) 1 = arcsin � 0,5 ∙ ∙ sin ( ) 1 � − Figure 2: Pressure angle over a full disc rotation Description Symbol Value Unit Center distance a 70.0 mm Normal pressure angle α n 22.5 ° Disc width b 10.0 mm Size ratio u 1.4 - Crowning c b 15.0 μm Arithmetic mean roughness Ra 0.1 μm Table 1: Specifications for the disc variant S70-A 0 30 60 90 120 150 180 0 500 1,000 1,500 contact pressure [MPa] 0 2 4 6 force [kN] contact pressure p H normal force F n φ A φ E φ C 2,000 8 F = 3 kN disc -10 -5 0 5 10 speed [m/ s] sum speed v Σ sliding speed v s φ A φ E φ C rotation angle [°] φ n = 1,000 rpm disc 0 1 2 3 4 disc width [mm] 0 1 2 3 4 rolling direction [mm] -4 -2 0 2 4 profile height [µm] φ A φ C φ E 0° 180° 0 30 60 90 120 150 180 + - Figure 3: Disc variant S70-A: Contact pressure and normal force (upper left) and sum speed and sliding speed (lower left) over half a disc rotation, disc sample (upper right) and representative section of an optical measurement of the disc surface at new condition (lower right) disc notation, S refers to the specimen, 70 represents the center distance, and A denotes the FZG test gears type A used in scuffing tests (hereafter referred to as type A gears). Because of the size ratio, the discs are of different sizes, with specimens categorized as S1 for the large disc and S2 for the small disc. For comparison, the parameters contact pressure, force, sum speed, and sliding speed are shown for both the S70-A discs and type A gears in Figure 3 and Figure 4. The size ratio of the discs enables the implementation of an uneven slide-to-roll ratio, with higher values in the area of positive sliding than negative sliding, based on the type A gear design. The contact pressure is likewise unevenly distributed. The disc surface features a 15 µm crowned finish and has been treated with tangential grinding, resulting in grinding grooves in the rolling direction, creating a uniform texture (Figure 3). The tooth flank surface of the type A gears exhibits the typical Maag 15 cross-hatch pattern (Figure 4). With an arithmetic mean roughness Ra of 0.1 µm measured in axial direction, the disc surface is smoother than the tooth flanks of the type A gears, which have an arithmetic mean roughness of 0.35 µm [DIN06]. 4 Experiments The scuffing load-carrying capacity tests are performed according to the FZG test procedures A/ 8.3/ 90 and A/ 16.6/ 90 on an FZG gear test machine, as specified in DIN ISO 14635-1 [DIN06]. Dip-lubrication is used, and the initial oil temperature of 90 °C is uncontrolled during the tests. Each test consists of twelve load stages including contact pressures from 146 to 1,841 MPa, with each stage lasting for 15 minutes (Table 2). The load increases after each completed load stage. The load stage at which scuffing damage occurs is referred to as the failure load stage. The disc tests were carried out on a two-disc tribometer. The testing procedure for the discs is in line with the FZG test procedures and includes the load stages four through twelve (Table 2). For each load stage, the test load on the two-disc tribometer was adjusted to ensure that the contact pressure at the pitch point of the discs matched that of the type A gears. The discs tests were conducted with injection lubrication and a controlled oil Science and Research 17 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0014 -0.5 0 0.5 1.0 1.5 0 500 1,000 1,500 contact pressure [MPa] 0 2 4 6 force [kN] 2,000 8 -10 -5 0 5 10 meshing coordinate [-] A C B D E -0.5 0 0.5 1.0 1.5 sum speed v Σ sliding speed v s speed [m/ s] T = 94 Nm pinion n = 1,450 rpm pinion normal force F n contact pressure p H - + 0 1 2 3 4 disc width [mm] 0 1 2 3 4 rolling direction [mm] -4 -2 0 2 4 profile height [µm] A B E Figure 4: Type A gears: Contact pressure and normal force (upper left) and sum speed and sliding speed (lower left) during meshing, gear sample (upper right) and representative section of an optical measurement of the tooth flank surface at new condition (lower right) Description Type A gears Discs S70-A Unit Test duration per load stage 15 15 min Load stages 1 - 12 4 - 12 - Contact pressure at pitch point 146 - 1,841 621 - 1,841 MPa Sum speed at pitch point 6.6 | 13.2 6.6 | 13.2 m/ s Rotational speed 1,500 | 3,000 1,000 | 2,000 rpm Oil temperature 90 (test start) 60 °C Lubrication Dip-lubrication Injection lubr. - *Gear test with SLC-1 : injection lubrication with a controlled oil temperature of 60 °C Table 2: Test parameters for gear and disc tests torque in the shafts. Lubricant is injected into each disc pair through a nozzle positioned at the midpoint of the center distance 5. The oil injection temperature and flow rate are monitored. 5 Experimental results on scuffing load-carrying capacity The tests revealed that the employed procedures cause scuffing damage to the discs. The analogy between the disc tests and gear tests is illustrated by the failure load stages (Figure 6). Tests conducted at higher speeds resulted in a lower failure load stage for both disc and gear tests. Using test oil SLC-H, both setups exhibited no scuffing with a failure load stage exceeding twelve. The disc tests show a consistent deviation of approximately two failure load stages, corresponding to around 300 MPa, compared to the gear tests. For this test plan, the gear test with SLC-L1 does not show a significant influence of the oil temperature and lubrication on the scuffing load-carrying capacity as demonstrated by the failure load stages. Science and Research 18 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0014 temperature of 60 °C, due to the system’s configuration. The test oils are categorized based on their scuffing loadcarrying capacity (SLC): two oils with low (L1 and L2), one oil with medium (M), and one oil with high (H) scuffing load-carrying capacity. The gear test with test oil SLC-L1 was carried out with a controlled oil temperature of 60 °C and injection lubrication, in order to determine the influence of these conditions on the scuffing load-carrying capacity during the tests. On the two-disc tribometer, the discs are mounted at the ends of the shafts, with each disc pair rotated 180° relative to the other 1 (Figure 5). The discs are driven by a synchronization gearbox with a gear ratio of -1 2. The required test force, measured by a load cell, is applied through a hydraulic cylinder and transmitted to the discs via a yoke and bearings 3. Acceleration sensors are attached to the yoke’s rocker arms. The disc force acts toward the centers of rotation and is supported at the contact point by the normal force, which bypasses the rotational axis, thereby generating torque 4. The double-disc arrangement compensates for the alternating Figure 6: Overview of the failure load stages from the gear and disc tests for all test oils Figure 5: Two-disc tribometer and depiction of a mounted disc pair The absolute sliding speeds of the discs are slightly lower than these of the type A gears due to the design and show a slide-to-roll ratio of 80 % at the damage point. Since the current setup of the two-disc tribometer does not yet allow for friction measurements, the friction-independent pv value is used to evaluate the scuffing load-carrying capacity on a tribological level [DYS75]. The pv value describes the contact-specific power and is determined by the product of the highest contact pressure and corresponding sliding speed at the point of scuffing. For both disc and gear tests, the pv values remain consistent across different test oils, irrespective of the test setup (Figure 7). 5.1 Damage analysis The scuffing damage on the disc exhibits a damage pattern similar to that observed on gears, with scuffing marks aligned with the rolling direction and resulting in material removal (Figure 8). The damage is evenly distributed across the disc surface but does not cover the entire width of the surface because of the crowning. Regarding the shape, the damage starts narrowly, widens to its maximum point, and then tapers slightly. Due to the test procedure, the damaged area on the disc appears more roughened since tests on the two-disc tribometer are terminated immediately after scuffing occurs, which is clearly indicated by an increase in vibration speed. As a result, the disc surface remains more worn compared to the gear surface. The damaged surface on the gear extends over two-thirds of the tooth flank in the rolling direction and across the full width of the active tooth flank (Figure 9). In both setups, the scuffing damage occurred in areas with the highest contact pressures and high positive sliding speeds. Science and Research 19 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0014 Figure 7: pv diagram of the damage points from the gear and disc tests for all test oils 3 disc width [mm] -20 -10 0 10 20 profile height [µm] 0 0 3 rolling direction [mm] 1 2 1 2  = 13.2 m/ s v Σ,C  failure load stage 8  p = 1,285 MPa H,max  test oil SLC-L1 Test parameters 0 500 1,000 1,500 contact pressure [MPa] 0 2 4 6 force [kN] φ A φ E φ C 2,000 8 -20 -10 0 10 20 speed [m/ s] rotation angle [°] φ φ A φ C φ E ω 2 v t 0 30 60 90 120 150 180 0 30 60 90 120 150 180 contact pressure p H normal force F n sum speed v Σ sliding speed v s Figure 8: Scuffing damage on a disc and a representative section of an optical measurement of the disc surface after testing (left), contact pressure and normal force (upper right) and sum speed and sliding speed (lower right) over half a disc rotation analogy to gear meshing in tests, as the tribological contact conditions of gears can be more accurately replicated. The two-disc tribometer, combined with the use of non-circular discs, provides an effective test method for determining the scuffing load-carrying capacity based on FZG test procedures. The disc tests produced comparable scuffing damage and demonstrated a strong correlation with the gear tests. The higher failure load stages Science and Research 20 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0014 6 Conclusion The non-circular discs enable the implementation of locally varying slide-to-roll ratios. The geometric design provides continuous and smooth meshing while forming a contour with a variable disc radius. Besides, the disc contour partially consists of curvature radii ratios that correspond to those of involute gears. This improves the -0.5 0 0.5 1.0 1.5 0 500 1,000 1,500 contact pressure [MPa] 0 2 4 6 force [kN] 2,000 8 -20 -10 0 10 20 speed [m/ s] meshing coordinate [-] A C B D E -0.5 0 0.5 1.0 1.5 v t 3 tooth width [mm] -10 -5 0 5 10 profile height [µm] 0 0 3 rolling direction [mm] 1 2 1 2  = 13.2 m/ s v Σ,C  p = 823 MPa H,max  test oil SLC-L1  failure load stage 5 Test parameters E B A normal force F n contact pressure p H sum speed v Σ sliding speed v s Figure 9: Scuffing damage on a type A gear and a representative section of an optical measurement of the gear surface after testing (left), contact pressure and normal force (upper right) and sum speed and sliding speed (lower right) during meshing Nomenclature A, B, D, E - Meshing points C - Pitch point F disc N Disc force F n N Normal force F test N Test force K - Contact point O - Center of rotation Ra µm Arithmetic mean roughness SRR - Slide-to-roll ratio T pinion Nm Torque pinion V oil l/ min Oil volume flow a mm Center distance b mm Disc width c b µm Crowning n disc rpm Rotational speed disc n pinion rpm Rotational speed pinion p H MPa Contact pressure p H,max MPa Maximal contact pressure pv kW/ mm 2 Contact-specific power r mm Contact radius r b mm Base circle radius r min mm Minimal disc radius u - Size ratio v s m/ s Sliding speed v t m/ s Tangential speed v Σ m/ s Sum speed v Σ,C m/ s Sum speed at pitch point α ° Pressure angle α n ° Normal pressure angle δ ° Auxiliary angle ξ ° Disc angle φ ° Rotation angle ω °/ s Angular speed 1, 2 Indices for large, small disc DIN German Institute for Standardization ISO International Organization for Standardization SLC Scuffing load-carrying capacity observed in the disc tests can be attributed to the lower sliding speeds compared to the type A gears. Therefore, a higher contact pressure is required to cause scuffing on the discs. Another aspect are the larger curvature radii of the discs which leads to a lower contact temperature at the same friction power due to the larger contact area. Moreover, the disc surface is smoother due to the grinding process which influences the friction and lubrication behavior. On a tribological level, the friction-independent pv value shows a strong correlation between the disc and gear tests. For further investigations, the setup of the two-disc tribometer should be upgraded to include friction and temperature measurement capabilities in order to determine friction-dependent parameters such as flash temperature. Furthermore, this would facilitate a more precise thermal analysis of the disc contact, as both the bulk temperature and the coefficient of friction can be used to determine the contact temperature. Literature [ALN04] Alanou, M. P.; Evans, H. P.; Snidle, R. W.: “Effect of different surface treatments and coatings on the scuffing performance of hardened steel discs at very high sliding speeds”. In: Tribology International 37 (2004). [ANG91] Angeles, J.; López-Cajún, C. S.: “Optimization of Cam Mechanics”. 1. Auflage. Springer Dordrecht, 1991. [AUF25] Aufderstroth, N, Schierholz, L., Vorgerd, J., Oehler, M.: “Modelling of non-circular discs as test specimens for gear analogy tests”. In: Proceedings of the Institution of Mechanical Engineers, Part J: Journal of Engineering Tribology (2025) [BRE17] Brecher, C.; Löpenhaus, C.; Mevissen, D.: „Zwei- Scheiben-Tribometer mit variablem Schlupfverlauf“. In: Antriebstechnik 12/ 2016 (2016). [CIH23] Cihak-Bayr, U.; Wopelka, T.; Wintersteiger, C.: “Investigation of friction and scuffing during loss of lubrication on a high velocity twin-disc test rig”. In: Forschung im Ingenieurwesen 87 (2023). [CON23] Contreras Urgiles, R. W.; Echávarri Otero, J.; Chacón Tanarro, E.; Franco Martínez, F.; Cortada-García, M.: “A test for evaluating the scuffing performance of fully-formulated lubricants”. In: Tribology International 187 (2023). [CZI20] Czichos, H.; Habig, K.-H.: “Tribologie-Handbuch: Tribometrie, Tribomaterialien, Tribotechnik”. 5., überarbeitete und erweiterte Auflage. Springer Vieweg, 2020. [DIN06] DIN ISO 14635-1: „Zahnräder - FZG-Prüfverfahren - Teil 1: FZG-Prüfverfahren A/ 8,3/ 90 zur Bestimmung der relativen Fresstragfähigkeit von Schmierölen (ISO 14635-1: 2000)“. 2006. [DIN15] DIN 51350-1: „Prüfung von Schmierstoffen - Prüfung im Vierkugel-Apparat - Teil 1: Allgemeine Arbeitsgrundlagen“. 2015. [DYS75] Dyson, A.: “Scuffing - a review: Part 2: The mechanism of scuffing”. In: Tribology International 3 (1975). [GRE22] Grenet de Bechillon, N.; Touret, T.; Cavoret, J.; Changenet, C.; Ville, F.: “A new experimental methodology to asses gear scuffing initiation”. In: Tribology - Materials, Surfaces & Interfaces 16 (2022). [ING15] Ingram, M.; Hamer, C.; Spikes, H.: “A new scuffing test using contra-rotation”. In: Wear 328 - 329 (2015). [LI13] Li, S.; Kahraman, A.; Anderson, N.; Wedeven, L. D.: “A model to predict scuffing failures of a ballon-disc contact”. In: Tribology International 60 (2013). [LIN22] Linke, H.; Börner, J.: „Stirnradverzahnungen: Berechnung - Werkstoffe - Fertigung“. 3., aktualisierte Auflage. Hanser Verlag, 2022. [NIE03] Niemann, G.; Winter, H.: „Maschinenelemente 2: Getriebe allgemein, Zahnradgetriebe - Grundlagen, Stirnradgetriebe“. 2. Auflage. Springer, 2003. [OPT24] Optimol Instruments: 2disk - Zwei-Scheiben-Prüfstand. [online] URL: https: / / optimol-instruments.de/ de/ produkte/ 2disk/ (Stand: 19.08.2024). [PAT95] Patching, M. J.; Kweh, C. C.; Evans, H. P.; Snidle, R. W.: “Conditions for Scuffing Failure of Ground and Superfinished Steel Disks at High Sliding Speeds Using a Gas Turbine Engine Oil”. In: Journal of Tribology 117 (1995). [PCS24] PCS Instruments: MTM [online] URL: https: / / pcs-instruments.com/ product/ mtm/ (Stand: 19.08.2023). [ROT04] Rothbart, H. A., Hrsg.: “Cam Design Handbook”. McGraw-Hill, 2004. [SAV17] Savolainen, M.; Lehtovaara, A.: “An experimental approach for investigating scuffing initiation due to overload cycles with a twin-disc test device”. In: Tribology International 109 (2017). [SCH24] Schierholz, L.; Aufderstroth, N.; Vorgerd, J.: „Analogieuntersuchungen an unrunden Zahnscheiben mit tribologisch äquivalenten Lastbedingungen wie im Zahneingriff“. In: Dresdner Maschinenelemente Kolloquium (2024). [SOM18] Sommer, K.; Heinz, R.; Schöfer, J.: „Verschleiß metallischer Werkstoffe: Erscheinungsformen sicher beurteilen“. 3. Auflage. Springer Vieweg, 2018. [TEN16] Tenberge, P.; Weibring, M.; Gondecki, L. (2016) Modellprüfstand (DE 10 2016 015 529.9). Deutsches Patent- und Markenamt. URL: https: / / register.dpma.de/ DPMAregister/ pat/ register? AKZ=1020160155299&CURSOR=0. [TEN22] Tenberge, P.; Vorgerd, J.; Gondecki, L.: “2-disc tribometer for various tests on sliding/ rolling contacts with tribological loads such as in tooth flank contacts”. In: VDI-Berichte Nr. 2389 (2022). [VOR23] Vorgerd, J.: “Wirkungsgrad und Fresstragfähigkeit schnelllaufender Stirnradverzahnungen mit chemisch glattgeschliffenen Oberflächen”. Dissertation Ruhr-Universität Bochum, 2023. Science and Research 21 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0014 Pries [P1] provided an approach for the holistic modelling of transient, semi-elliptical contact areas, including head, root and edge load-bearing. Bünder [B2] complemented the load distribution model with a comprehensive representation of the contact forces. Neugebauer [N1] extended the design by adding criteria for mixed friction states and analysed the influence of deflection angle and temporal variability of the friction coefficient on typical damage patterns. Spura [S2] developed a new holistic simulation model for load distribution and investigated the wear behaviour of crowned, displaceable spline couplings. Science and Research 22 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 1 Introduction The reliability and service life of displaceable spline couplings determine the safety of many drive systems, from industrial gearboxes to wind turbines. Their characteristic crowned tooth flanks compensate for shaft misalignment, but they give rise to an oscillating Hertzian contact between a cylinder and a plane with complex frictional behaviour, in which hydrodynamic, mixed and solid-body friction alternate. Until now, design approaches have largely been confined to static press-fit analyses or to empirical metrics for individual operating conditions, while a method for early quantification of linear wear loss under varying contact conditions has been lacking. This paper links Archard’s phenomenological wear law [A1] with Kragelski’s molecularmechanical fatigue theory [K1] and Fleischer’s energybased fundamental equation [F1]. Instead of a constant wear coefficient, the reciprocal of the wear energy density derived from the elastic contact deformation is employed. Validation by means of test-rig experiments and field measurements demonstrates very good agreement with the observed wear values and provides a robust basis for the engineering design and optimisation of displaceable spline couplings. 2 State of the Art Investigations by Benkler [B1] into load distribution in toothed couplings form the basis for research on displaceable spline couplings. Heinz [H1] extended this work to the tooth kinematics under deflection, introduced the parameters pressure overlap and load-carrying number, and adapted the flash temperature hypothesis for the determination of frictional energy and tooth contact temperature. Strauß [S1] studied run-in and wear behaviour under variable operating conditions and defined a critical contact temperature as the threshold for adhesion. Wear energy density for wear prediction of displaceable spline couplings Christian Spura* Presented at GfT Conference 2025 This study presents a comprehensive predictive methodology combining the Archard wear model, Kragelski’s molecular-mechanical fatigue theory, and Fleischer’s energetic wear framework to accurately forecast linear wear depth in displaceable spline couplings. The specific crowned tooth flank geometry generates a cylinder-to-plane contact accompanied by oscillatory relative motion, which complicates traditional wear evaluation methods. To address this, the wear energy density is introduced as a key parameter, enabling a mechanistically founded quantification of wear progression. Extensive validation through dedicated test rig experiments and field trials confirms the high reliability of the model, with deviations between predicted and observed wear remaining below 9 % for flank pressures up to 900 N/ mm 2 . The presented approach facilitates early-stage wear prediction during the design process, supporting effective optimisation measures to enhance component durability and service life. Furthermore, the model’s adaptability to varying lubrication conditions and material pairings underscores its practical relevance for engineering applications involving complex contact scenarios. Keywords displaceable spline couplings, gear couplings, gear shaft connections, wear energy density, wear prediction calculation, molecular-mechanical fatigue theory, energetic wear fundamental equation Abstract * Prof. Dr.-Ing. Christian Spura Orcid-ID: https: / / orcid.org/ 0000-0001-8307-8919 FH Münster Department of Mechanical Engineering Stegerwaldstraße 39 48565 Steinfurt Archard [A1] developed the well known phenomenological wear model. Kragelski [K1] considered wear as a fatigue process at the molecular level, in which cyclic deformations lead to material removal. Fleischer [F1] supplemented these approaches with an energy balance and defined the wear energy density. Despite their different approaches, all three models can be transformed into one another. 3 Simulation model and load distribution The load distribution is determined using the simulation model presented in [S2], which was developed for crowned tooth flanks. The basis is an analytical stiffness model that represents the deformations of external and internal gearing and the influence of the tooth edge. FEA-based studies provided correction factors to improve the simulation model. Real gear deviations can be taken into account to produce a realistic flank contact pressure distribution. The results provide a reliable basis for the design and optimisation of displaceable spline couplings. Figure 1 shows, as an example, the flank contact pressure distribution along the tooth width of an ideal gear according to DIN 5480 (52 × 2 × 30 × 24). Case (a) has a crowning radius of 80 mm and a deflection of 2 degrees. Case (b) has a crowning radius of 1200 mm and a deflection of 0.1 degrees. As a result of the kinematics, the contact point moves in both the tooth width and the tooth height directions, and the displacement in the tooth width direction is decisive for the loading and the wear behaviour. 4 Wear energy density The following provides a comparative discussion of the energy-based wear fundamental equation [F1, F2] and the fundamental equation of fatigue theory [K1, K2]. Both approaches are based on the assumption of a critical number of contacts and permit a description of the linear wear intensity I h for a single contact. With reference to [F3], the term frictional energy density e R is replaced in the following by the term wear energy density e V , with wear number v v = 1.0. The energy-based wear intensity is given by the ratio of frictional work to wear energy density: (1) with: friction coefficient μ, mean contact pressure p a , and wear energy density e V . Under the assumption of a run-in contact, that is equilibrium of the surface roughnesses, the fatigue theory describes the linear wear intensity I h by (2) and the associated minimum friction coefficient by (3) where C 1 is a dimensionless material factor, α H is the hysteresis coefficient, τ 0 is the shear resistance under the prevailing lubrication conditions, t is the fatigue exponent (typical range 2 to 12), k is a coefficient that depends on the internal stress distribution and on the material (typical range 3 to 7), σ B is the tensile strength, E red is the reduced elastic modulus, and β is a piezo coefficient. Equating (1) and (2) and using (3), then rearranging, yields the wear energy density (4) h = 1 ∙ H ∙ a ∙ 0 2 ⁄ ∙ � ∙ B � ∙ � 1 red � 1− 2 ⁄ = � 0 ∙ H red � 1 2 ⁄ + h = ∙ a V Science and Research 23 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 Figure 1: Flank contact pressure distribution of an ideal gear (DIN 5480: 52 × 2 × 30 × 24). a) Crowning radius 80 mm at 2° deflection; b) Crowning radius 1200 mm at 0.1° deflection. V = B 1 ∙ ∙ H ∙ 0 2 ⁄ ∙ � 1 red � ( 1− 2 ⁄ ) ∙ �� 0 ∙ H red � 1 2 ⁄ + � ( −1 ) The energetic wear model proposed by Fleischer [F1, F2], as expressed in Equation (12), relates the frictional work W R (the product of friction force F R and sliding distance s g ) to the wear volume V V via the wear energy density e V . The wear volume V V may also be determined from the nominal contact area A 0 and the linear wear depth h V . (12) A comparison of Equations (11) and (12) clearly indicates that the wear coefficient k H may be interpreted as the reciprocal of the wear energy density e V . Consequently, k H can be directly obtained from the energetic contact parameters, thereby obviating the need for a separate determination of the wear coefficient through complex tribological testing in model applications. (13) 5.1 Operational wear Based on the calculation of the wear energy density e V as given in Equation (4), the linear wear depth in the low-wear stage of the i-th tooth per revolution can be determined using Equations (12) and (13) as follows: (14) In this context, both the linear wear depth h V and the wear energy density e V are to be determined separately for the shaft tooth and the hub tooth. 5.2 Initial wear The running-in phase is typically characterised by a distinctly higher wear rate. Both the duration of this phase and the magnitude of the wear depend on the specific loading conditions and may vary significantly between different tribological systems. During the running-in process, smoothing of rough surface regions and adaptation of the macro-geometry take place. As a result, the corresponding load distribution can be approximated iteratively. According to [S2], a modified calculation based on Equation (15) is introduced to determine the linear wear depth during the running-in phase, h VE (tooth thickness reduction). This extension of Equation (14) incorporates, in addition to the previously defined parameters, the running-in time t VE (10 min to 40 min), the rotational speed n, and a dimensionless correction factor f VE (range: 10 2 to 10 4 ) to account for the increased wear rate resulting from the lower effective wear energy density: (15) Here too, the linear wear depth h VE (tooth thickness reduction) in the running-in phase is to be determined separately for the shaft tooth and the hub tooth. R = R ∙ g = ∙ N ∙ g = V ∙ V = V ∙ 0 ∙ ℎ V H = V ℎ V , i = i ∙ N , i ∙ g V ∙ 0 , i ℎ VE , i = VE ∙ i ∙ N , i ∙ g ∙ ∙ VE V ∙ 0 , i Science and Research 24 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 Thus the wear energy density can be determined from the physical and mechanical material properties, the shear resistance, the hysteresis coefficient and the loading. The remaining quantities are defined as follows: (5) (6) (7) (8) (9) with: Young’s modulus E, Poisson’s ratio ν, and flank contact pressure σ Fl . The energy-based wear fundamental equation [F1] relates the wear energy density e V and the linear wear intensity I h via the frictional shear stress τ R : (10) Note that the calculation of the wear energy density e V and of the linear wear intensity I h must be carried out separately for each of the two contacting bodies. 5 Wear prediction calculation It should be noted that the present calculation refers exclusively to the low-wear stage and does not account for the subsequent progressive high-wear stage. Initial wear can, in practice, be quantified; however, it results from several superimposed wear mechanisms, making an unambiguous attribution of the causes impracticable. The Archard wear model [A1, A2], given in Equation (11), describes the volumetric wear V V as a function of the normal force F N , the sliding distance s g , and the wear coefficient k. Depending on the model variant, the hardness H of the material is either explicitly included in the denominator of Equation (11) or already incorporated within the wear coefficient k H . The ranges of k or k H vary considerably with the underlying tribological system and must be established by means of experimental investigation. (11) red = 1 1 − 12 1 + 1 − 22 2 1 = 0,12 ∙ 16 �2∙ 5 � 2,6 � −5 4� = 1,5 ∙ � 4 ∙ ( 1 − − 2 ) + ( 1 − 2 ∙ ) 2 2 = − Fl ∙ H 1,45 ∙ red 0 = Fl 2 ∙ H 2,1 ∙ red ℎ = = ∙ Fl V = ∙ N ∙ g = H ∙ N ∙ g 5.3 Wear compensation calculation Using Equations (14) and (15) to determine the linear wear depth in the low-wear stage as well as during the running-in phase, an iterative wear compensation calculation can be performed. For each tooth pair, the linear material loss is successively added to the existing single pitch deviations, and on this basis, the resulting load distribution is recalculated iteratively. This process is first carried out for the running-in phase, followed by the operational wear in the low-wear stage. To adequately represent the running-in dynamics, five to ten iterations are usually sufficient. Running-in wear leads, on the one hand, to a reduction in tooth thickness and thus to a change in gear stiffness, and, on the other hand, compensates for pitch errors. Both effects improve the load distribution and consequently reduce flank pressure. Furthermore, according to [D1, K3], the compensation of pitch errors through running-in wear improves gear quality by one to two tolerance grades. Similar mechanisms apply to operational wear, although with significantly less impact than during the running-in phase. Figure 2 illustrates the normalised flank pressure distributions for two different types of misalignment-tolerant spline teeth. Figure 2a shows the results for a gear coupling with a tooth profile according to DIN ISO 21771 (analogous to a running gear profile). In this case, the tall and slender tooth profile provides high compliance, enabling pitch errors along the tooth circumference to be compensated and ensuring an even load distribution across all tooth pairs. The maximum flank pressure remains comparatively low even when individual pitch deviations are taken into account. Figure 2b presents the pressure distributions for a spline connection with a tooth profile according to DIN 5480. The short and stiff tooth profile leads to a more pronounced load concentration. This imbalance causes increased running-in wear in the initial stage. As a result of the running-in wear, the tooth flanks adapt to the actual load, increasing the number of load-carrying tooth pairs from 21 to 24. Consequently, the maximum pressure decreases to about 68 % of its original value, and the load distribution becomes noticeably more uniform. In the gear coupling, running-in wear is less pronounced, as all available tooth pairs are already involved in load sharing before the running-in phase. As a result, flank pressure is reduced to around 94 % of the initial value. This confirms, from a computational standpoint, that running-in wear achieves a compensation of pitch errors equivalent to an improvement in gear quality by one to two tolerance grades, as described in [D1, K3]. 5.4 Wear-related service life prediction The wear-related service life can be determined directly from the maximum permissible linear wear depth h V,zul , the applied loads, the wear energy density e V , and a sta- Science and Research 25 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 Figure 2: Normalised flank pressure distribution of the splines with module m = 2 mm, number of teeth z = 24, crowned circle radius r B = 1200 mm, angular misalignment ε = 0.07°, torque T = 1700 Nm, and quality grades 8 (shaft) and 10 (hub). Tooth profile according to: a) DIN ISO 21771; b) DIN 5480. The shaft and hub materials employed were 18CrNi- Mo7 - 6 and 42CrMo4, respectively. A closed-loop oil circulation system, equipped with filtration and temperature monitoring, ensured consistent and reproducible lubrication conditions. Three lubricants of identical viscosity were tested: a mineral oil, a polyalphaolefin, and an ester-based oil. The experimental objectives encompassed the verification of load-bearing capacity criteria related to pitting and scoring resistance, as well as the determination of safe threshold values for flank pressure and mixed friction regimes. Furthermore, the influence of pressure overlap on lubricant supply and resultant wear was investigated, supplementary load-bearing criteria were identified, and the predicted linear wear depth in the low-wear stage was validated. Additional studies addressed the impact of lubricant type, maintaining constant viscosity, and facilitated the classification of observed wear mechanisms. 6.1 Results of experimental investigations and wear prediction calculations The test points were selected to encompass a broad spectrum of flank pressure, specific frictional power, and sliding velocity, as illustrated in Figure 3. A region characterised by pronounced low-wear stage was utilised for the direct validation of the model predictions. Test specimens exhibiting consistent mild wear and maintaining the low-wear stage were classified as successful runs, whereas specimens without a low-wear stage or showing signs of scoring, pitting, and flank breakouts were categorised as failures. Science and Research 26 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 tistical correction factor. In the basic formula presented in [F4]: (16) the proportion of frictional energy α 1,2 is explicitly specified, and the maximum sliding velocity v g,max is used. As shown in [S2], α 1,2 may be omitted if e V is determined using Equation (4) and v g,max is replaced by the sliding distance per revolution s g and the rotational speed n. This yields the service life L h , which is to be calculated separately for each counterbody (shaft or hub): (17) with: quantile C x (for 90 % survival probability = -1.28; for 50 % = 0), coefficient of variation C v (range for steel-steel pairings: 0.1 to 0.6). 6 Experimental investigations The validation of the wear prediction model was conducted on a machining test rig designed in accordance with DIN 51354, featuring a maximum torsional moment of 1750 Nm, a maximum rotational speed of 1500 min −1 , and a deflection angle capability of up to 2°. The setup comprises two test shafts, each equipped with two independently adjustable gearing sequences, enabling the simultaneous evaluation of four distinct geometry configurations per test cycle. Deflection angles are precisely controlled by a mechanical adjustment mechanism. h = 10 −3 ∙ V ∙ �ℎ V , zul − ℎ VE � R ∙ g ∙ ∙ ( 1 − x ∙ v ) ∙ 60 + VE = V ∙ �ℎ V , zul − ℎ VE � 3600 ∙ 1 , 2 ∙ R ∙ � g , max � ∙ ( 1 − x ∙ v ) + VE h Figure 3: Results of the experimental investigations. The normalised experimental results for the linear wear depth, presented in Figure 4, demonstrate the high accuracy of the computational models introduced in Sections 4 and 5. For enhanced comparability, all measured values were normalised to the respective maximum wear depth. Deviations of up to 9 % were observed under very high flank pressures in the range of 500 N/ mm 2 to 916 N/ mm 2 , which is considered acceptable given the extreme load conditions. For flank pressures typically encountered in practical applications, up to 500 N/ mm 2 , the discrepancy between experimental data and model predictions is 6 % or less, thereby confirming the excellent predictive capability of the wear prognosis calculation. The parameters utilised for the wear prediction calculations are listed in Table 1. 6.2 Effects of pressure overlap The pressure overlap factor λ σ is a key parameter characterising the lubrication condition in the tooth contact. It describes the ratio between the Hertzian contact width 2a and the contact path length s k along the tooth flank, thereby providing information on lubricant supply and the development of the hydrodynamic lubricating film. If λ σ < 1.0, the contact path fully covers the contact width, which promotes the formation of a load-bearing lubricating film. For λ σ > 1.0, an overlap of contact regions occurs, potentially impairing lubricant supply. Under such conditions, both material stress and adhesion tendencies increase, leading to local mixed or boundary lubrication regimes that promote scuffing. The characteristic W-shaped wear profile appears at λ σ < 1, with the highest wear located at the reversal point of the contact motion, as shown in Figure 5a. At these reversal points, the lubricating film is absent, resulting in solid contact with abrupt transitions between static and sliding friction. Even at low sliding speeds below approximately 100 mm/ s, scuffing initiates due to micro-welds producing rough particles that accelerate wear. At elevated temperatures above approximately 150 °C, scuffing may also occur, causing severe damage to the tooth flanks. If λ σ > 1.0, a U-shaped wear profile develops, as illustrated in Figure 5b. A permanent region of the tooth flanks remains in continuous contact, which restricts lubricant supply and only permits a limited lubricating film. Mixed Science and Research 27 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 Figure 4: Comparison of normalised linear wear depth h V in the low-wear stage between calculated predictions and experimental results. Material 18CrNiMo7-6 750 0,02 7,93 0,20 18CrNiMo7-6 (case-hardened) 1150 0,02 7,90 0,24 42CrMo4 800 0,02 8,04 0,20 42CrMo4 (nitrided) 1000 0,02 8,00 0,22 Table 1: Parameters used for the wear prediction calculations. Figure 6. Despite differing wear profiles (U-shaped vs. W-shaped), the values of wear energy density and linear wear intensity within the respective low-wear stages fall within the same range. 6.4 Wear phenomena and damage patterns The following section describes and characterises the wear mechanisms and associated damage patterns observed on the test rig. The results correspond closely to those reported from field tests. Science and Research 28 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 friction alternates with areas of intensified solid contact, and the maximum wear is concentrated in the overlap zone. Scuffing also occurs here, as the material surface is subjected to persistently high shear and adhesion forces. 6.3 Wear energy density and linear wear intensity The experimental results reveal distinct running-in and operational regions for the wear energy density e V and the corresponding linear wear intensity I h , as shown in Figure 5: Wear profile as a function of the pressure overlap factor: a) λ σ < 1.0, b) λ σ > 1.0. Figure 6: Ranges of wear energy density e V and linear wear intensity I h for the running-in and operational regions (low-wear stage). • Scoring: Linear grooves oriented along the sliding direction over the tooth width. They result from abrasive loading when asperities of the harder partner, hardened wear particles, or hard foreign particles penetrate the softer contact partner. These grooves are often superimposed with local plastic indentations. • Indentations: Appear as shallow depressions frequently surrounded by ring-shaped material buildup at their edges. They arise from the impact of flat particles (contaminants or wear debris) that do not embed into the substrate. Indentations often coincide with scoring and are difficult to detect macroscopically; thus, SEM or optical microscopy is recommended for reliable identification. • Scuffing: Characterised by sharply defined material tears bordered by fine scoring patterns. It occurs under insufficient lubrication and high flank pressures when welding of contact points happens and subsequent sliding causes material removal. • Pitting: Small, circular depressions formed by particle accumulation on the contact surface combined with minor relative motion. They are frequently accompanied by plastic deformation or scoring at their edges due to locally elevated contact pressures. • Scars, dimples, ripples (vibrational wear): Manifest in three closely related forms: scars (associated with irregular vibration amplitude), dimples (from twodimensional oscillation), and ripples (from one-dimensional oscillation). Common features include local accumulation of wear debris resulting in irregular surface patches, circular depressions, or periodic transverse structures. Transitions and mixed forms among these types are frequently observed. • Shear cracks: Linear microcracks oriented transverse to the sliding direction, sometimes branching. They develop when the local shear strength of the material is exceeded due to pointwise overload and are mainly found near scuffing sites or scuffing grooves. • Surface fragmentation: Under oscillatory motion with small deflection angles, fine crack networks and severely notch-inducing fragmentation occur. The combination of high local stresses and insufficient debris removal leads to a rugged surface topography often accompanied by abrasive marks. • Worm tracks: Appear as elongated breakouts perpendicular to the sliding direction, often displaying characteristic discoloration. They result from adhesive welding at elevated temperatures and insufficient lubrication, where repeated sliding tears local welds and fatigue progressively forms breakout structures. Microstructural analyses often reveal tribo-martensite and temperature peaks up to approximately 800 °C. • Profile alteration: If wear progresses continuously after the running-in phase, macroscopic steps and material buildup form on the tooth profile. This wear progression exceeds permissible profile tolerances and may ultimately lead to plastic deformation or tooth fracture. Countermeasures generally include: (a) optimisation of lubricant selection (CLP oils with EP additives, high viscosity), (b) minimisation of lubrication deficiencies and particle contamination, (c) use of appropriate materials and heat treatments (nitriding, case hardening), (d) reduction of contact pressures and vibration amplitudes, and (e) avoidance of steel-on-steel contact via surface coatings. In summary, a combination of design measures and appropriate lubrication has proven to be an effective approach to counteract the complex and superimposed wear mechanisms in misalignment-tolerant spline couplings. 7 Conclusions The presented results demonstrate that the combined model of Archard, Kragelski, and Fleischer provides a reliable prediction of linear wear depth even under varying contact conditions, thereby offering a robust basis for design modifications and service life optimisation. The model was validated for flank pressures up to 900 N/ mm 2 and commonly used lubricants. Mean deviations between simulation and experimental data are 6 % for flank pressures up to 500 N/ mm 2 and 9 % up to 900 N/ mm 2 . Future work will focus on investigating additional lubricant types and viscosities as well as alternative material pairings. The integration of real-time data from condition monitoring systems could further enhance the model’s accuracy. In conclusion, the combined model delivers a dependable approach for the early quantification of linear wear depth and makes a valuable contribution to the structural optimisation of displaceable spline couplings. Literature [A1] J. F. Archard: Contact and Rubbing of Flat Surfaces. Journal of Applied Physics, 24 (8), 1953, S. 981-988. https: / / doi.org/ 10.1063/ 1.1721448 [A2] J. F. Archard, W. Hirst: The wear of metals under unlubricated conditions. Proceedings of the Royal Society of London. Series A, Mathematical and physical sciences, 236 (1206), 1956, S. 397-410. https: / / doi.org/ 10.1098/ rspa.1956.0144 [B1] Benkler, H.: Der Mechanismus der Lastverteilung an bogenverzahnten Zahnkupplungen. Diss., TH Darmstadt, 1970. [B2] Bünder, C.: Analyse der Beanspruchungen der Verzahnungen von Zahnkupplungen. Diss., TU Dresden, 2000. Science and Research 29 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 [K1] I. W. Kragelski: Reibung und Verschleiss. VEB Verlag Technik Berlin, 1968. [K2] I. W. Kragelski, G. Fleischer, W. S. Kombalov, U. Winkelmann: Vereinigung der Ermüdungstheorie und des energetischen Ansatzes zur Berechnung des Verschleißes. Schmierungstechnik, 10 (5), 1979, S. 132-136. ISSN 0036-6226 [K3] Kollmann, F. G.: Welle-Nabe-Verbindungen - Gestaltung, Auslegung, Auswahl. Springer-Verlag, Berlin Heidelberg New York, 1984. ISBN 3-540-12215-X [N1] Neugebauer, H.: Verzahnungsbeanspruchbarkeit. FVA Forschungsvorhaben Nr. 307/ II, FVA-Heft 712, Frankfurt a. M., 2003. [P1] Pries, M.: Geometrie und Kinematik von Bogenzahnkupplungen. Diss., TU Dresden, 1991. [S1] Strauß, E.: Einsatzgrenzen und Einlaufverhalten von nichtgehärteten Zahnkupplungen. Diss., TH Darmstadt, 1984. [S2] Spura, C.: Tragfähigkeitsberechnung und Verschleißanalyse von bombierten Zahnwellenverbindungen. Dissertation, RWTH Aachen, 2012. Science and Research 30 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0015 [D1] Dietz, P.: Die Berechnung von Zahn- und Keilwellenverbindungen. Erschienen im Selbstverlag des Verfassers, Büttelborn, 1978. [F1] G. Fleischer: Energetische Methode der Bestimmung des Verschleißes. Schmierungstechnik, 4 (9), 1973, S. 269- 274. ISSN 0036-6226 [F2] G. Fleischer: Energiebilanzierung der Festkörperreibung als Grundlage zur energetischen Verschleißberechnung. Teil 1: Schmierungstechnik, 7 (8), 1976, S. 225-230, Teil 2: Schmierungstechnik, 7 (9), 1976, S. 271-279, Teil 3: Schmierungstechnik, 8 (2), 1977, S. 49-58. ISSN 0036-6226 [F3] Fleischer, G.: 40 Jahre Bewertung von Reibung und Verschleiß mit Hilfe der Energiedichte. Tribologie und Schmierungstechnik, 51. Jahrgang, 3/ 2004, S. 5-11. ISSN 0036-6218 [F4] Fleischer, G.: Bereitstellung tribologischer Kennwerte von Reibpaarungen zur Beurteilung und Prognose der Lebensdauer. Schmierungstechnik, 12 (1981) 8, S. 233- 235. ISSN 0036-6218 [H1] Heinz, R.: Untersuchung der Kraft- und Reibungsverhältnisse in Zahnkupplungen für grosse Leistungen. Diss., TH Darmstadt, 1977. State of the art and motivation In industrial gear systems, a balanced lubricant system is essential for ensuring the longest possible service life of the components. Radial shaft seals (RSS) can be used to reliably seal the gear system. Made of an elastomer material, the RSS’s are subjected to thermal, chemical, and mechanical stress during operation, which can lead to different wear and aging characteristics at the sealing edges. Depending on their severity, these wear and aging characteristics can impair the sealing ability, making safe operation of the transmission system no longer possible. In order to investigate the compatibility of a sealing system - consisting of RSS, lubricant, and shaft - dynamic sealing ring tests are often carried out in industry (SEW63190052.07, 2024). The wear and aging characteristics are evaluated visually and haptically according to the state of the art in order to assess the compatibility of the lubricant and elastomer. The damage characteristics evaluated can be divided into chemical, physical-thermal, and thermal primary damage (Bauer, 2021). This paper examines the physical-thermal damage characteristics of “hardening” and “hard deposits”. A high seal edge temperature can lead to post-crosslinking of the elastomer and thus to the aging characteristic of “hardening” (Bauer, 2021). On the model test bench for dynamic testing of RSS, the lubricants can degenerate due to a high oil sump temperature and test duration. This can cause “hard deposits” to form on the sealing edge surface, consisting of aged lubricant and elastomer residues. The aging characteristics of an RSS, as well as the “hard deposits” on the sealing edge, are evaluated according to (SEW63190052.07, 2024) with a rating system. As already shown in the studies (Alt, et al., 2023) and (Wilbs, et al., 2023), nanoindentation can be used to characterize the wear and aging of the elastomer at the sealing edge due to µ-mechanical material changes. By determining suitable parameters for characterizing the damage, the analysis can be carried out in a measurable and reproducible manner in the future. Furthermore, the evaluation of the stiffness of the tribologically stressed elastomer via the indentation depth according to (Wilbs, et al., 2023) enables a more detailed analysis of the properties of the deposits on the sealing edge, which are collectively referred to as “hard deposits” according to (SEW63190052.07, 2024). The aim of the present study is to calculate the stiffness discretely over the indentation path and to examine the characteristics of the stiffness gradient. Mechanical characterization The aim of mechanical characterization is to determine mechanical parameters for stressed RSS’s that can be correlated with the previous haptic characterization. For Science and Research 31 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0016 Systematic investigation of the µ-mechanical material change of the sealing edge of radial shaft seals Felix Bernhardt, Katrin Alt, Markus Wöppermann* Presented at GfT Conference 2025 The stability of sealing systems is tested on dynamic model tests. The wear and aging characteristic are evaluated haptically and optically according to the state of the art to rate the lubricant-elastomer compatibility. A nanoindenter can be used to investigate µ-mechanical material changes because of tribological stress in radial shaft seals. The results of new measurement methods are correlated with the haptic evaluations of the sealing edges and allow for distinction in the tribologically induced layer on the sealing edge. Keywords radial shaft seal, dynamic test, µ-mechanical characterization, nanoindentation, elastomer-lubricant compatibility Abstract * M.Sc. Felix Bernhardt M.Sc. Katrin Alt Dr.-Ing. Markus Wöppermann SEW-EURODRIVE GmbH & Co KG Ernst-Blickle-Str. 42 76646 Bruchsal/ Germany the sealing edge as standard, and the mean value and standard deviation are then calculated. During post-processing, the dissipated work W diss and the reversible work W rev are calculated from the hysteresis when the force is applied over the indentation path, as shown in Figure 1. The dissipated work W diss describes the proportion of material damping, while the reversible work W rev describes the proportion of reversible deformation work during deformation reversal due to unloading. The relative dissipated work rel.W diss is calculated as the relative magnitude of material damping compared to elasticity, which represents the ratio of dissipated work W diss to total work W ges = W diss + W rev . The maximum force F max is recorded at the maximum indentation depth. The stiffness curve over the indentation depth is evaluated for each measuring point up to the maximum indentation depth. Results When plotting the relative dissipated work over the maximum force according to (Wilbs, et al., 2023) in Figure 2, clusters are formed from an internal data set of haptically evaluated RSS’s that were tribologically stressed. The color coding of the clusters indicates the corresponding degree of hardness based on the tactile assessment. The tribologically non-stressed RSS’s form a cluster at the lowest relative dissipated work rel.W diss and maximum force F max . The maximum force F max increases proportionally to the degree of hardening from the haptic evaluation. As hardening increases, the relatively dissipated work also increases up to the point of slight hardening, and then decreases with increasing maximum force F max up to a point of significant hardening. If the stiffness is evaluated incrementally over the indentation depth up to the maximum indentation according to (Wilbs, et al., 2023), two different characteristic stiffness curves result Science and Research 32 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0016 the tests with the nanoindenter from the study (Alt, et al., 2023), a probe tip with a radius of 40 µm at the spherical hemisphere of 30° is used. In the study (Alt et al., 2023), the sealing lips of the RSS are turned inside out for vertical indentation on the sealing edge. One measurement method is distance-controlled quasi-static indentation followed by de-indentation (retraction) of the probe tip. The distance control sets the appropriate indentation depth to reduce friction of the elastomer material if the indentation depth is too great. The maximum indentation depth is -60 µm and both the indentation rate and the unloading rate are 5 µm / s. The preload force is 1,91 mN. For each RSS, 20 measurement points are recorded on Figure 1: Calculation of dissipated work W diss , reversible work W rev , and maximum force F max . Figure 2: Cluster of the correlation between the haptic evaluation and the relative dissipated work over the maximum force. from the examination of RSS with haptically evaluated deposits. An example of a sealing edge with the two characteristic stiffness curves per measuring point and as an average value can be seen in Figure 3. The upper sealing edge shows a deposit with a stiffness curve that rises and then falls to the maximum indentation depth at the individual measuring points. The sealing edge shown below exhibits high initial stiffness at the start of the indentation depth, which initially decreases rapidly over a short indentation path. Subsequently, the mean value shows an approximately constant stiffness curve up to the maximum indentation depth. The measurement points in this area show a spread between small amounts of positive and negative gradients. Discussion Figure 2 shows a clear cluster structure of the degrees of hardening based on the haptic evaluation and the measured values of the relatively dissipated work rel.W diss over the maximum force F max . This allows initial limit values in the relative dissipated work rel.W diss and the maximum force F max , which correlate with the haptic evaluation and thus enable the hardening to be characterized on the basis of measureable parameters. It is also clear that the dispersion per haptic evaluation category is high and that future aging characterization based on measured values is necessary. The examples in Figure 3 show that there are two different characteristic curves of stiffness over the indentation depth, which, according to (SEW63190052.07, 2024), can only be evaluated with a defined damage characteristic of “hard deposits”. This indicates the need for differentiation in the characteristic “hard deposits”. The generic term “tribologically induced layer formation” is introduced, which can manifest itself in a “deposit” or “hard layer”. Figure 4 shows the schematic layer structures of both tribologically induced layers based on optical observation in connection with the measured stiffness curves over the indentation depth. Science and Research 33 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0016 Figure 3: Stiffness curves per measuring point and average value for two sample sealing edges with the same rating of a significant “hard deposit” with the grade 3 according to (SEW63190052.07, 2024). Figure 4: Schematic representation of the structure of tribologically induced layers and the corresponding stiffness curve over the indentation depth. Conclusion As part of this work, a fixed number of RSS examples are examined using a nanoindenter to define limit values in mechanical characteristics that correlate with previous haptic assessments and could replace them in the future, enabling characterization based purely on measured values. The previous term “hard deposit” according to (SEW63190052.07, 2024) is generally understood. Based on the findings of the different prevailing stiffness curves, the tribologically induced layers as a general term can be differentiated into “hard layers” or “deposits” and evaluated measurably. References Alt Katrin [et al.] Measurement device and automation solution for analysing tribologically damaged radial shaft seals / / Reibung, Schmierung und Verschleiß, GfT e.V.. - 2023. Bauer Frank Federvorgespannte-Elastomer-Radial-Wellendichtungen: Grundlagen der Tribologie & Dichtungstechnik, Funktion und Schadensanalyse. - Wiesbaden, Deutschland : Springer-Verlag, 2021. SEW63190052.07 SEW 63190052.07 Prüfvorschrift: Statische und dynamische Prüfungen von Radialwellendichtringen (RWDR). - [s.l.] : SEW-EURODRIVE GmbH & Co KG, 2024. Wilbs Christian [et al.] µ-Mechanical characterization of tribologically stressed elastomer surfaces with respect to radial shaft sealing systems / / Reibung, Schmierung und Verschleiß, GfT e.V.. - 2023. Science and Research 34 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0016 On the left side of Figure 4, is the stiffness curve of the “deposit” with the increase and subsequent decrease in stiffness over the indentation depth. When looking at the associated sealing edge, a viscous mass consisting of lubricant and elastomer components can be seen. After indentation, a probe tip impression clearly shows that this deposit is irreversibly deformed. The stiffness is therefore low at the beginning of the indentation. Underneath this is intact elastomer material, which is usually hardened in the upper layer due to the high temperature. When the probe tip hits the hardened elastomer material, the stiffness is higher. This explains the maximum stiffness. As the indentation depth increases further and more surrounding material is deformed, the stiffness decreases up to the maximum indentation depth and the hysteresis shows a degressive stiffness curve. This degressive stiffness curve is typical for hardened elastomer material. The right-hand side of Figure 4 shows the stiffness curve of the “hard layer” with high stiffness at the beginning and a rapid decrease followed by slight positive gradients up to the maximum indentation depth at a sample measuring point. Visual and tactile examination of the associated sealing edge reveals a thin, brittle, and cracked layer. This explains the very high stiffness at the beginning of the indentation. The rapid drop in stiffness is related to the probe tip breaking through the hard layer. The subsequent slight increase, which accounts for a maximum of 10 % of the change in stiffness of the stiffness decrease, is due to the indentation of elastomeric material without hardening under the layer. EP-Additives Extreme pressure (EP) additives are critical components in modern metalworking fluids, designed to protect tools and workpieces under the most demanding conditions. These additives play a vital role in machining operations such as cutting, drilling, milling, and forming, where high loads, temperatures, and friction are common. Without EP additives, tool wear would increase significantly, surface finishes would degrade, and productivity would suffer. EP additives are chemical compounds formulated to react with metal surfaces under high pressure and temperature conditions. Their primary function is to form a protective film on metal surfaces, preventing direct metal-to-metal contact. This film reduces friction and wear, especially in boundary lubrication regimes where the lubricant film is too thin to fully separate the surfaces. Common EP additives include chlorinated paraffins (CLP) and sulfur based compounds (sulfur carriers). Phosphorus-based compounds are also often referred to as EP additives, but mainly work as antiwear additives. Chlorinated paraffins CLPs are saturated, polychlorinated hydrocarbons with a chain length of approximately 10 to 30 carbon atoms and a chlorine content of about 35 to 70 %. For nearly 100 years, chlorinated paraffins have been used as EP additives in metalworking fluids to reduce or prevent adhesive wear, especially during metal processing. Chlorinated paraffins are classified according to the number of carbon atoms in their molecules: short-chain (SCCPs, C 10 -C 13 ), medium-chain (MCCPs, C 14 -C 17 ), long-chain (LCCPs, C 18 -C 20 ), and very long-chain (VCCPs, C> 20 ). Chlorinated paraffins are poorly biodegradable; SCCPs and MCCPs are acutely and chronically very toxic to aquatic organisms, bioaccumulative, and persistent in the environment. They accumulate in the food chain of humans and animals and can now be detected in water, soil, sewage sludge, as well as in many living organisms, in human fatty tissue, and breast milk. Since SCCPs are also potentially carcinogenic and toxic to reproduction, their use in metalworking fluids has been banned in the EU since 2002. Current efforts, such as those by ECHA and the UN Stockholm Convention, aim to heavily restrict the use of medium-chain chlorinated paraffins as well. There are also intentions to ban or restrict the use of MCCPs by e.g. the South Korean and Japanese authorities and by Health Canada. In the longer term, such measures are also expected for long-chain chlorinated paraffins. A corresponding assessment of very longchain chlorinated paraffins is not foreseeable at this time. However, due to their very high viscosity and limited solubility in base oils, these compounds are only conditionally suitable for use as additives in metalworking fluids. In many negative substance lists used in the manufacturing industry, chlorinated paraffins are only permitted in metalworking fluids when technically necessary (e.g., in stainless steel forming). Science and Research 35 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0017 EP Additives with Enhanced Sustainability for Water-miscible Metalworking Fluids Wilhelm Rehbein, Isabell Lange, Kevin DiNicola, Salvatore Rea, John Williams* Presented at GfT Conference 2025 Extreme pressure additives are essential components for many water-borne metalworking fluids. They are used to prevent adhesive wear and cold welding by generating protective layers on metal surfaces under boundary lubrication conditions in heavy duty cutting and forming processes. Sulfur carriers are high-performing EP-additives that can support the metalworking industry on their way to a more sustainable future. Keywords Extreme Pressure additives, Sulfur carriers, Chlorinated paraffins, Water-miscible metalworking fluids, Sustainability, Wear prevention Abstract * Dipl.-Ing. Wilhelm Rehbein Isabell Lange LANXESS Deutschland GmbH, Mannheim, Germany Salvatore Rea Kevin DiNicola John Williams LANXESS Corporation, Naugatuck, CT, USA tain between 1 and 5 sulfur atoms, which form a bridge connecting the hydrocarbon or ester structures of the molecules. Similar to chlorinated paraffins, they form a protective layer of metal sulfides that prevents cold welding between the tool and the workpiece. Due to their higher polarity compared to sulfurized hydrocarbons, sulfurized esters can also reduce friction even under low mechanical or thermal stress. Sulfur carriers are available with various molecular structures, differing in polarity and activity, allowing them to be precisely tailored to the requirements of specific metalworking processes. Their performance spans a wide temperature range - from slow machining operations to high-speed processing. Their effectiveness can be further enhanced through synergistic combinations with overbased calcium sulfonates, polycarboxylates, or other polar compounds. In machining processes, such as deep hole drilling, sulfur carriers prevent the formation of long chips that could damage or block the tool. This is achieved by forming metal sulfide layers on the chip surfaces, which make the chips more brittle and cause them to break more easily. Modern sulfur carriers are light in color, have low odor, and are compatible with most additives used in lubricants. No special measures are required for their disposal. In terms of human and environmental toxicity, they are significantly more favorable than chlorinated paraffins. Unlike chlorinated paraffins, sulfur carriers are not subject to labeling requirements under GHS or hazardous substance regulations. Many sulfur carriers are more than 50 % derived from renewable raw materials, such as vegetable oils or vegetable oil methyl esters. Lots of them are even suitable for Science and Research 36 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0017 Chlorinated paraffins in metalworking processes CLPs are used in metalworking coolants as extreme pressure (EP) additives, meaning they prevent adhesive tool wear that occurs under high mechanical pressure. In the first step, the polar chlorinated paraffin molecules adhere to the metal surface, forming a friction-reducing film. In a second step, under sufficient mechanical or thermal stress, a chemical reaction between the chlorine atoms and the metal surface forms a very thin layer of metal chlorides (see Figure 1). Due to its salt-like structure, this layer is significantly less shear-resistant than the metal itself. As a result, cold welding or adhesion of the workpiece material to the tool is prevented. This would otherwise lead to the breaking out of larger particles from the tool surface. In metalworking, chlorinated paraffins are highly effective at low machining speeds. At higher machining speeds, which usually also lead to increased temperatures at the tool and workpiece, they decompose and form hydrogen chloride, which causes increased tool wear. In the presence of water or high humidity, there is a risk of hydrolysis. The hydrochloric acid formed in this process corrodes tools, workpieces, and the entire machine tool. Compared to other EP additives, the price of chlorinated paraffins is relatively low. However, disposal costs are high in many regions, as they must be incinerated in special high-temperature combustion facilities to prevent the formation of toxic dioxins. Sulfur carriers As an alternative to chlorinated paraffins, sulfurized hydrocarbons and sulfurized synthetic or natural esters are used as EP additives. These so-called sulfur carriers con- Figure 1: Formation of adsorption and reaction layers by EP-additives use in lubricants that meet the requirements of the EU Ecolabel and the US Vessel Incidental Discharge Act. Sulfur carriers can be easily emulsified and are suitable components for soluble oils and semisynthetic metalworking fluids. Some sulfur carriers are even water-soluble and work as excellent EP additives for synthetic cutting and forming fluids. Tribological test results All sulfur carriers in the following tests can be certified as: ▪ based on renewable raw materials for more than 50 % ▪ having a low impact on the environment ▪ not labelled as hazardous to humans or to the environment ▪ meeting the requirements of the LuSC list for environmentally acceptable lubricants To evaluate the performance of the sulfur carriers with improved sustainability in comparison with chlorinated paraffins in emulsifiable metalworking fluids, tribological tests were conducted under laboratory conditions. Pin & Vee tests The Pin & Vee Test is a widely recognized laboratory method used to evaluate the extreme pressure (EP) performance of metalworking lubricants. It simulates the conditions under which lubricants need to prevent wear, seizure, and cold welding between metal surfaces under high load and low speed which are typical in machining operations such as drilling, milling, and turning. The test apparatus consists of two stationary V-shaped blocks (the “Vees”) and a rotating cylindrical pin. The Vee block is typically made of hardened steel and features two angled surfaces that cradle the pin. The pin, also made of steel, rotates under a controlled load and speed while being pressed into the Vee block. The lubricant under test is applied to the contact area between the pin and the Vee block. During the test, a progressively increasing load is applied to the Vees until a failure condition is reached, usually indicated by a sudden rise in friction, temperature, or visible damage to the metal surfaces. Pin & Vee test results of metalworking fluid formulations can be easily transferred to the performance of the formulations in metalworking processes such as tread cutting or thread forming. When running the tests with different metalworking fluid emulsions, the following parameters were used: ▪ All tests were done with emulsions of soluble oil concentrates, containing 40 % mineral oil and 10 % EPadditive. The soluble oils were diluted by 2 % in deionized water. ▪ The test Pins were made of SAE 3135 carbon steel, the Vee blocks are made of AISI 1137 carbon steel ▪ The tests were carried out acc. to ASTM D3233A with a speed of rotation of 290 +/ - 10 rpm and a temperature at test start of 22 °C. After a run-in phase of 5 minutes at 300 lbs, the load was continuously increased from 300 to 4,500 lbs. The EP-performance of three different sulfur carriers and one medium chain chlorinated paraffin was compared in the Pin & Vee test by using these parameters (Table 1). When comparing the different sulfur carriers by Pin & Vee test (Figure 2), the highest load of > 4,500 lbs was achieved by the formulation containing 10 % inactive sulfurized triglyceride. This result shows the excellent friction reducing properties of this additive. With 10 % active sulfurized ester, the torque curve over time was slightly lower but the test stopped due to the break of the Pin at approximately 4,200 lbs. The reason Science and Research 37 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0017 EP additive Chemical structure Sulfur resp. chlorine content Kin. Viscosity at 40°C Content of sustainable raw material MCCP Medium chain chlorinated paraffin, C 14 -C 17 50% Cl 180 mm²/ s 0% Active sulf. ester Sulfurized methyl ester 17% S (8% active S) 55 mm²/ s > 80% Inactive sulf. tg Sulfurized vegetable triglyceride 9.5% S (<1% active S) 230 mm²/ s > 80% Active sulf. tg Sulfurized vegetable triglyceride 18% S (9% active S) 220 mm²/ s > 80% Table 1: properties of the EP-additives tested by Pin & Vee and four-ball tests Science and Research 38 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0017 Figure 3: Pin & Vee test results of soluble oil formulations diluted to 2 % with deionized water. The formulations contain no EP-additive respectively a medium chain chlorinated paraffin or an active sulfurized triglyceride. Figure 2: Pin & Vee test results of 3 different sulfur carriers. The dashed lines show the applied loads, the solid lines the resulting torques. for this effect is that a protective layer was formed already at lower mechanical load and temperature due to the higher activity of this sulfur carrier. However, because of the lower viscosity and polarity of this sulfur carrier, the protective layer was less stable and broke down at the maximum load of 4,200 lbs. The small torque peak at the maximum load indicates the collapse of the protective layer, following by a direct metal-to-metal contact and cold welding. The test result with the 10 % active sulfurized triglyceride is even slightly lower with a maximum achievable load of approximately 4,100 lbs. The lower maximum achievable load is probably caused by a kind of competitive situation between the protective layer which is formed by the active sulfur, and the physical adsorption layer formed by the polar centers of the sulfurized triglyceride molecule. Compared with the test result of the diluted soluble oil formulation containing 10 % of the MCCP or with the maximum achievable torque of the diluted soluble oil which does not contain any EP-additive, even the active sulfurized triglyceride showed a significantly better performance (Figure 3). Four-Ball tests In addition to the Pin & Vee tests, Four-Ball tests were carried out to evaluate the capability of the EP-additives to prevent abrasive and adhesive wear under high load and high speed conditions. The test results are transferable to real cutting processes with rotating tools or work pieces, e.g. turning or milling processes. The ASTM D2783 Four-Ball extreme pressure test evaluates the EP properties of lubricating fluids. It helps differentiate between fluids with low, medium, and high EP performance, which is crucial for applications involving high-load conditions. In this test, three steel balls are clamped together in a cradle, and a fourth ball is rotated against them under increasing loads. ▪ In this test, the speed of rotation is 1800 rpm. The tests starts at ambient temperature (approx.70 °F / 21 °C) and runs for 10 seconds. Additional to the weld load, which is the load at which the rotating ball welds to the stationary balls, there are two more parameters: ▪ The last non-seizure load gives an indication about the stability of the protective layer, formed by the lubricant additives, at increasing loads. It is the highest load at which no seizure (metal-to-metal contact without lubrication) occurs during the test. The higher the last non-seizure load, the better the lubricant is suited for high-pressure applications ▪ The load-wear index (LWI) is a measure of a lubricant’s ability to withstand extreme pressure conditions. It indicates how well a lubricant can prevent wear as the load increases. The LWI is calculated from test results at various load levels where wear occurs. It’s a weighted average of the loads at which wear is observed. A higher LWI value means better lubricating performance and wear protection. The ASTM D4172 Four-Ball wear test is used to assess the wear preventive characteristics of lubricating fluids Science and Research 39 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0017 Figure 4: Four-Ball test results of soluble oil formulations containing no respectively different EP-additives according to ASTM D2783 and ASTM D4172. major concerns when using sulfurized additives in water miscible metal working fluids are: ▪ The release of H 2 S due to a decomposition of the sulfur chain. ▪ An increase of the acidity of the water phase due to decomposition of the sulfur carrier, either at the sulfur chain or at the ester group. The hydrolytic stability of the sulfur carriers was tested in a modified ASTM D2619 Beverage Bottle test (Figure 5). To simulate the chemical conditions of a waterborne metalworking fluid, the pH of the water phase was adjusted to a value of 9.0. In this test, 75 mL of a blend of 20 % sulfur carrier and 80 % mineral oil and 25 mL of water, adjusted to pH 9.0 Science and Research 40 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0017 under sliding contact conditions and indicates the lubricants ability to prevent abrasive wear. Similar to D2783, three stationary steel balls are clamped together, and a fourth ball is rotated against them. ▪ The parameters in these tests are a speed of rotation of 1200 rpm, a starting temperature of 86 °F (30 °C), a test duration of 60 minutes and a constant load of 20 kg. Again the different EP-additives were tested with 10 % concentration in soluble oil formulations which were diluted to 10 % with deionized water. The results of the Four-Ball EP test (Figure 4) impressively show that sulfur carriers are significantly more effective in preventing adhesive wear when compared to medium chain chlorinated paraffins. The weld load of the inactive sulfurized triglyceride is on the same level as the MCCP, the weld loads of the more active sulfurized ester and sulfurized triglyceride are higher. Even more, the last non-seizure load and the load-wear-index, showing the stability of the protective layer and the overall resistance to adhesive wear, are much higher for the sulfur carriers. The same trend is visible in the ASTM D4172 wear tests. Without any EP-additive, it was not possible to run the test due to strong friction vibrations. With the different sulfur carriers, the Four-Ball wear is almost on the same level and approximately 30 % lower than with the chlorinated paraffin. Hydrolytic stability For water-miscible metalworking fluids, the hydrolytic stability of their components plays a crucial role. The Figure 5: Beverage Bottle test, used for the evaluation of the hydrolytic stability of the sulfur carriers Table 2: test results of the beverage bottle test of the sulfur carriers. The results outlined in red are the most important ones. were filled in a glass bottle with a volume of approx.. 207 mL (7 fl.oz). The bottle is sealed and placed in a rotating oven which is heated up to 93 °C (200 °F). After 48 hours of continuous rotation, the liquids are filled in a separatory funnel and the acid number of both the oil and water phases are measured (Table 2). After 48 hours in the Beverage Bottle tests, all three sulfur carriers show no change of the acidity of the water phase or oil phase and also almost no release of H 2 S. This is proof that sulfur carriers are stable even under the conditions in a water-mixed cooling lubricant and are not prone to hydrolysis. Summary The test results demonstrate that it is possible to formulate water-miscible metalworking fluids with excellent performance by using EP additives based on sulfurized renewable raw materials. All sulfur carriers in these tests can be certified as: ▪ Based on renewable raw materials for more than 50 %. ▪ Having a low impact on the environment. ▪ Are not labelled as hazardous to humans or to the environment. ▪ Meeting the requirements of the LuSC list for environmentally acceptable lubricants. The Pin & Vee test results show that sulfur carriers with improved sustainability have a positive influence on energy consumption and tool life in cutting and forming processes when used as EP additives in metalworking fluid emulsions. Higher Four-Ball weld loads, last non-seizure loads and load-wear-indices and lower Four-Ball wear test results indicate that these sulfur carriers can effectively prevent adhesive and abrasive wear in water-borne metalworking fluids and exceed the performance of conventional additives like chlorinated paraffins. Science and Research 41 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0017 Science and Research 42 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0018 Introduction The vision of “Tribometry 4.0” is an application-oriented, data-driven tribology that, beyond individual parameters, enables a deep understanding of dynamic interactions in the tribological system [GREB23, GREB23a]. Test rigs are evolving into digital platforms that combine flexible test planning, networked data acquisition, and simulation-supported analyses [HEINL25]. The goal of modern “Tribometry 4.0” is to reproduce tribological systems in the laboratory in an applicationoriented manner so as to understand relationships and interactions - and not, as was often customary in the past, to generate simple key figures for glossy brochures. Background The need for new test methods and evaluation methods arises from the numerous challenges currently facing technology in general and mechanical engineering in particular: a) Replacement of well-known and proven raw materials (lead, SAPS, metal-containing additives, PFAS); use of recycled materials and raw materials This results in the necessity of numerous tests to evaluate new materials/ chemicals in order to find substitutes during the development phase and, ultimately, to release them prior to market launch. Rapid, cost-effective, yet still meaningful tests are therefore required (so-called “high-throughput testing”). b) Focus on energy efficiency and friction reduction If one wishes to measure low friction down to super-lubricity, highly accurate and reproducible friction measu- Modern Application-Oriented Tribometry: Understanding Tribology Instead of Producing Characteristic Values - How Modern Measurement Technology Can Enable a View into the Hidden Tribological Contact Markus Grebe, Henrik Buse, Richard Heinlein, Andreas Keller* Presented at GfT Conference 2025 Application-oriented, modern tribological testing technology (“Tribometry 4.0”) is a decisive development tool on the way to a high-performance and reliable product. Based on tribological system analysis and the testing strategy derived from it, the potential of various optimization approaches can only be investigated and evaluated within an acceptable framework in terms of both time and cost by using meaningful laboratory tests. However, the current disruptive developments in many areas of automotive and mechanical engineering require completely new test rigs, test methods, and evaluation approaches adapted to the new requirements. This paper presents these new trends and the resulting requirements and testing technology in detail. Using the example of four laboratory tests for the high-speed suitability of lubricating greases, the evaluation of the performance of plain bearings, and investigations into the presence of grease during long sliding movements, it is shown how additional measurement variables and video documentation with automatic image evaluation help to better understand the processes involved in tribological contact. Keywords Tribometry 4.0, Advanced Measurement Techniques, Machine Learning, Lubrication States, Electric Drives, Image Analysis Abstract * Dr. Markus Grebe Orcid-ID: https: / / orcid.org/ 0000-0002-7048-2645 Dr. Henrik Buse, Andreas Keller Kompetenzzentrum Tribologie Mannheim (KTM) at Technische Hochschule Mannheim, 68163 Mannheim Richard Heinlein Tribologie - Engineering Mannheim GmbH (TEMa) Auf dem Hochschulcampus, 68163 Mannheim predestines them for thermal management (keyword “one-fluid” for thermal management and lubrication of mechanical components). c) Electric and alternative drives A disruptive development for automobile construction is the transition from combustion engines to electric drive systems. In this context, it is initially irrelevant whether the vehicles are battery-electric or fuel-cellpowered. This transition results in new focus areas in the mass market that previously only played a role in special-purpose machine building. Figure 2 provides a conrements are required. This poses major challenges for mechanics and metrology, because high normal forces have to be applied and measured while the acting friction forces (tangential forces) are only a fraction of these (lubricated systems in mixed lubrication approx. f = 0.12; hydrodynamics or super-lubricity f < 0.01 [BMWE25]). Friction measurement results must be critically reviewed with other measured variables. Under this heading, the increasing use of water-based lubricants can also be addressed, as these have low internal friction and thus allow low shear and churning losses. In addition, their heat capacity is high, which Science and Research 43 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0018 Figure 2: Challenges in the context of electric drive concepts Figure 1: Our definition of Tribometry 1.0 to 3.0 - 1.0: First described tribological test rig; 2.0: First commercially available tribometer distributed worldwide; 3.0: Standard friction, wear, and temperature measurement b) Digital Twin A major future topic will be the use of digital twins of tribometers. There is still great potential here. However, as is generally the case in tribometry, the challenge will remain in reconciling simulation with real application. c) Machine Learning The practical use of artificial intelligence - and specifically machine learning - in tribometry is still in its infancy. Online analysis of measurement signals (categorical recognition of operating states such as boundary friction, mixed friction, hydrodynamics, sudden failure/ seizure), prediction of the further course of tests, and improved interpretation of tests through multidimensional linking of measured variables are major topics for the near future. Numerous challenges must also be overcome here, such as generating “good” and FAIR data [GARA25], selecting suitable algorithms, and objectively validating the approaches and solutions. Concrete Application Examples for the Use of Additional Sensors, Variables, or Video Analysis Below, we present four concrete application examples from ongoing projects. a) Vibration Measurement on a High-Speed Rolling Bearing Test Rig As part of a laboratory study on the suitability of lubricating greases for high-speed applications (within FAM Science and Research 44 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0018 cise overview of this challenge. Widely discussed are problems with current passage and electrical discharge, which are being addressed in the FAM standards working group AA653 [NAK53]. High rotational speeds also lead to lubrication issues, which will be discussed in more detail later (application example a), as well as lifetime issues due to the resulting high number of cycles. In the context of electric drive systems, investigations into hydrogen as an energy carrier can also be considered, which likewise poses major challenges for tribology. From these challenges arise concrete requirements for new test rigs and the associated measurement technology. These are illustrated schematically in Figure 3. Digitalization in Tribology and Tribometry In addition to test rigs, digitalization in the field of tribology and tribometry will play an increasingly decisive role in the future. This affects three major areas: a) Metrology In order to increase the informative value of measurements, high-frequency data acquisition of as many variables as possible is necessary. Structure-borne sound or vibration measurements in particular require high sampling rates in the range of several kilohertz. Challenges then arise in data reduction and data storage. Current research is focusing more and more on automatic data analysis and feature engineering rather than simple reduction via mean values [HEINL24]. Figure 3: Requirements for test rigs and measurement technology NA 062-06-52 - Lubricating Greases), comparative investigations were carried out on four sample greases on various high-speed rolling bearing test rigs (including spindle bearing test rig WS22) [NAK52]. KTM took part in this study together with various industrial partners. A newly developed high-speed rolling bearing test rig of the institute was used (Test rig details in [GREB22]). In contrast to the other laboratories, KTM uses, in addition to classical friction torque and temperature measurement, high-resolution vibration measurements to detect signs of starvation and disturbances in the rolling kinematics at an early stage (Figure 4). The results show that, by means of vibration measurement and acoustic emissions (Figure 5, bottom row), critical lubrication states can be detected significantly earlier than with the classical variables, which often only respond when total failure has occurred. Science and Research 45 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0018 Figure 5: Speed ramp in the framework of the laboratory study within NA 062-06-52; top: speed ramps (red) and friction torque evolution (blue); middle: temperature trends at various locations of the bearings; bottom from left to right: acceleration, noise, and peak signals (noise) over load steps (grey) Figure 4: Temperature, acceleration, friction force and sound as measurable values c) Timeand Position-Resolved Contact Voltage Measurement to Determine Lubrication State In the FVA research project “Grease Presence” (FVA- 987), the friction and lubrication behavior in grease-lubricated sliding contacts with long stroke is being investigated [KELL25, KELL25a]. In the field of rolling bearings, important new findings on the lubrication behavior of greases are currently being obtained using optical tribometers and computer simulations [POLL19; WAND21; ZANG23]. Similar knowledge is still lacking for sliding contacts with long stroke, as found, for example, in linear guides or in screw drives. In the project, model tests are carried out in a pin-on-plate configuration (PoP) as well as original-component tests on trapezoidal screw drives of a rear-axle steering system. In the PoP tests, decreasing friction values are frequently observed over the sliding cycle (so-called “bathtub curve”), which could not be evaluated and explained with confidence [KELL24]. With the aid of a self-developed position-resolved contact voltage measurement, the lubrication state can now be correlated with the friction coefficient [KELL25a]. It is evident that hydrodynamic states are reached at times despite the comparatively low sliding speed of an oscillating motion and the lack of an ideal wedge geometry (Figure 7). These changes of the lubrication regime explain the low friction values and the favorable wear behavior. In Figure 8 one can see how the model system pin-onplate, initially lubricated with grease, first runs in (graphic at bottom right: hydrodynamic shares at mixed friction increase) and then relatively abruptly fails as a result of the sudden onset of starved lubrication (Figure 8, Science and Research 46 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0018 b) Vibration Measurement with ML Evaluation on a Sliding Bearing Test Rig Within the industrial working group “Machine Learning in Tribology”, in which five industrial companies joined forces, the question was investigated how the friction and wear behavior of a radial plain bearing can be examined more precisely, i.e., more informatively. For this purpose, in addition to the classical variables friction torque and temperature, a highly accurate online wear measurement, contact voltage measurement, and structure-borne sound measurements were used (Figure 6). It was again shown that the information content of highfrequency vibration measurements clearly exceeds that of the classical parameters and that the measurements are considerably more sensitive when it comes to describing system states [HEINL24]. First mixed-lubrication contacts could be detected early, repeatably, and with clear separation, whereas no excursions were yet visible in the friction and temperature trends. In combination with contact voltage measurement, this metrology allows reliable prediction of when the running-in of a bearing is complete and when critical operating states occur, for example as a result of overload or starved lubrication. The transferability of trained anomaly-detection models was also investigated with regard to bearing material and lubricant. The algorithm-based combination of run-in detection, definition of a baseline state, and detection of deviations from this state (anomaly detection) is a strategy that can also be applied to other test rigs. A current doctoral project is investigating whether the features and ML algorithms generated in the laboratory allow direct predictions for original components and aggregates in the field. Figure 6: Measurement data of a test with a run-in radial plain bearing and average temperature; third from top: wear signal (unfiltered and mean); bottom: acoustic emission signal (blue) and contact voltage signal (red) single graphic at bottom left: sudden rise in friction; collapse of the lubricant film at bottom right visible in the contact voltage signal). d) Grease Presence Assessment Using a Camera System and ML-Based Evaluation In the same project (FVA - Grease Presence), the influence of different lubricant depots that form on the moving friction partner and on the counterbody was to be in- Science and Research 47 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0018 Figure 8: Measurement and video data of a model test of oscillating sliding with long stroke; top left: friction coefficient over stroke cycle (friction hysteresis) (red) and contact voltage signal (black); top right: image documentation of the left and right end position (captured from time-lapse video); bottom left: friction trend over test time (peak (orange) and mean values (blue)); bottom right: temperature (blue) and change in the contact voltage signal over the test duration Figure 7: Friction coefficient (red) and contact voltage signal (black) over one friction cycle (hysteresis, forward and backward stroke ±15 mm) vestigated. For this purpose, a camera system was applied that observes the moving friction partner during the test. Using automated image evaluation, the grease quantities at the end positions and on the moving friction partner can now be determined automatically and user independent (Figure 8). In addition, it is possible to detect a color change of the grease. This generally correlates very well with the onset of wear due to starved lubrication, since even the In summary, the following can be stated once again: • Test rigs, test methods - and later also standards - are required that can simulate the new practical questions in the laboratory in an application-oriented manner. • Digital solutions such as “digital twin” and “machine learning” must be more strongly linked to test technology. • The goal of “Tribometry 4.0” must be to understand tribology and not to generate key figures for advertising brochures. The challenges offer great potential for further development and an increase in the significance of test technology. References [BMWE25] Information des Bundesministeriums für Wirtschaft und Energie; https: / / www.energiefor schung.de/ de/ glossar/ Supraschmierung+%28eng l%3A+superlubricity%29; Download 20.8.2025 [GARA25] Garabedian, N.T., Schreiber, P.J., Brandt, N. et al. Generating FAIR research data in experimental tribology. Sci Data 9, 315 (2022). https: / / doi.org/ 10.1038/ s41597-022-01429-9 [GREB22] M. Grebe: Influence of a vibration load on the service life of rolling bearings in e-drives; 7 th World Tribology Conference 2022, Lyon 10-15 2022 [GREB23] M. Grebe: Tribometry 4.0; NextLub Conference, Presentation and Digital Proceedings, Uniti/ FVA/ GfT, Düsseldorf, 2023. [GREB23a] M. Grebe: Tribology 4.0: From Da Vinci to digitalisation; Lube Magazine - The European Lubricants Industry Magazine; Issue 178, p. 15-17; December 2023 Science and Research 48 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0018 smallest amounts of abrasion lead to darkening (Figure 9). Within that project it was shown that the combination of contact voltage measurement and automated visual evaluation of grease presence and grease condition provides significant added value compared to the classical variables friction force and temperature. The question of how individual grease deposits contribute to lubricating contact could thus be answered for the tribosystem in question. Summary This publication has shown that tribometry is continuously evolving. Current topics such as electromobility or sustainability issues bring new challenges for tribological test technology. The combination with new technical possibilities forms the basis for the next stage in tribometry, which we refer to as “Tribometry 4.0”. Using four concrete application examples from ongoing research projects, it was demonstrated how additional measurement technology and AI-based evaluation methods help to provide insights into the tribological contact, which is usually hidden. They clearly show the added value of new variables or features and of image analysis compared to the classical friction and wear parameters. The methods presented, as well as further approaches, for example based on impedance measurement technology, are being continuously developed in further projects (e.g. [TIDO25]) and in the context of industrial research in order to understand tribology ever better in the future and to advance developments in a targeted and thus timeand cost-efficient manner. Figure 9: Video documentation with automated pixel evaluation [HEINL24] R. Heinlein, D. Glowania, G. Tidona, M. Grebe: Einsatz von maschinellem Lernen in der Schmierfett-Evaluierung; Vortrag GfT Jahrestagung 2024; Tagungsband Vortrag 69, S. 426-435, ISBN 978- 3-9817451-9-1; 2024. [HEINL25] R. Heinlein, M. Grebe: Expertensystem zur Erkennung des Schmierungszustandes auf Basis dynamischer Versuchsführung; Vortrag anlässlich des DVM-Workshops „Zuverlässigkeit tribologischer Systeme“, Mannheim, 13. März 2025 [KELL24] A. Keller, M. Grebe: Verbesserung des Verständnisses der Fettrückhalte- und Schmierungsmechanismen von oszillierenden Gleitkontakten mit großem Hub; Vortrag GfT Jahrestagung 2024; Tagungsband Vortrag 15, S. 95-102, ISBN 978-3- 9817451-9-1; 2024. [KELL25] A. Keller, M. Enhuber, M. Grebe, K.-H. Jacob: Abschlussbericht zum FVA-Eigenmittelvorhaben FVA 987-I: „(Vor-)Entwicklung einer Prüfstrategie zur Qualifizierung von Schmierfetten für lebensdauergeschmierte Gleitanwendungen mit langem Hub“; 2025 [KELL25a] A. Keller, D. Kursawe, M. Grebe, K.-H. Jacob: Zwischenbericht zum FVA-Eigenmittelvorhaben FVA 987-II: „Einfluss der Fettdegradation durch thermo-oxidative Alterung auf die Schmierung von Gleitanwendungen mit langem Hub“; 2025 [NAK52] Protokoll „NAK Drehzahlkennwertbestimmung“ im FAM NA 062-06-52 - Schmierfette; 05. Mai 2025 (WebCon) [NAK53] Protokollentwurf AA653_RV3 im „NAK Elektrische Kennwerte“ vom 17.7.2025 im FAM NA 062-06-53; (WebCon) [POLL19] G. Poll; X. Li; N. Bader; F. Guo: Starved Lubrication in Rolling Contacts - A Review; Bearing World Journal Vol. 4 (2019), pp. 69-81 [TIDO25] G. Tidona, M. Grebe: Zwischenbericht zum IGF- Forschungsprojekt „Einsatz von maschinellem Lernen in der Schmierfettentwicklung“ (DGMK 871), Industrielle Gemeinschaftsforschung (IGF); 2025 [WAND21] S. Wandel; N. Bader; F. Schwack; J. Glodowski; B. Lehnhardt; G. Poll: Starvation and Relubrication Mechanisms in Grease Lubricated Oscillating Bearings; Tribology International; 09/ 2021; https: / / doi.org/ 10.1016/ j.triboint.2021.107276 [ZANG23] S. Zhang; G. Jacobs; S. Vafaei; S. von Goeldel; F. König: CFD investigation of starvation behaviors in a grease lubricated EHL rolling contact; Forschung im Ingenieurwesen, 2023; https: / / doi.org/ 10.1007/ s10010-023-00633-2 Science and Research 49 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0018 the non-covalent bonds between the molecules of the thickener are broken, which shortens the fiber lengths and irreversibly degrades them physically. This causes the grease to soften and change its rheological properties. (Meijer, et al., 2023). This study examines the effect of total physical thickener failure in a grease in a real gear application. It shows that the irreversible shearing of the thickener cannot be simulated using established model tests for grease. Based on this, a modified model test is developed that replicates the effect. The modified model test developed is used as a method for formulating an adapted thickener system of the same type. The suitability of the modified model test for early-stage development is validated in a real gearbox application. Physical thickener degradation in planetary gearboxes The investigation of the degradation effect is based on the results of (Ochs, et al., 2024), which showed, among other things, that a defined gear test after just 300 hours of operation allows a significant differentiation of the condition of fundamentally suitable greases and that the Science and Research 50 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0019 Introduction In industrial gear systems, the lubrication system can be designed as a grease-lubricated system for cost-driven applications or to reduce the risk of leakage. Greases for gears generally consist of one or more base oils, a thickener, and various additives. Metal soap-based thickener systems are widely used in industrial applications and impart the viscoelastic properties of gear grease through their fibrous network. The subgroup of complex soaps, which are composed of a base, a fatty acid, and a non-fatty acid, are characterized by their high performance. (Kuhn, 2017). This study examines greases based on complex soap, which are primarily intended to provide lifetime lubrication in planetary gears in precision applications. The lubricity and service life of a grease are limited by physical and chemical degradation mechanisms caused by shear stress, pressure, high variations in operating conditions, and temperatures in the gear. Chemical degradation includes additive degradation and oxidation reactions at elevated temperatures, which are accompanied by bleeding and evaporation of the base oil. Physical degradation is primarily caused by the shear stress and shear rate exerted, which leads to increased oil separation and destruction of the thickener structure. Both types of degradation are irreversible processes (Rezasoltani, et al., 2016). At moderate operating temperatures (40 °C -70 °C), physical degradation is one of the most dominant damage processes that shortens the remaining service life of the grease (Rezasoltani, et al., 2016). The energy introduced into the system by shear stress dissipates into heat, and Design and Implementation of a Model Test for Reproducing Thickener Degradation in Grease-Lubricated Gearboxes and Its Role in Grease Formulation Development Katrin Alt, Markus Wöppermann, Frank Plenert, Jürgen Liebrecht* Presented at GfT Conference 2025 The irreversible structural degradation of thickeners in lubricating greases under gearbox operating conditions cannot currently be replicated using conventional model test methods. To overcome this limitation, a modified model test has been developed that enables the reproduction of this failure mechanism. This modified model test is used as a screening test for the formulation of alternative lubricating greases, and their suitability is validated in gearbox tests. Keywords Lubricating grease, Thickener degradation, Gearbox, Model test, Shear stress, Grease formulation Abstract * M.Sc. Katrin Alt Dr.-Ing. Markus Wöppermann SEW-EURODRIVE GmbH & Co KG Ernst-Blickle-Straße 42, 76646 Bruchsal Dipl.-Ing. (FH) Frank Plenert Dr.-Ing. Jürgen Liebrecht FUCHS LUBRICANTS GERMANY GmbH Friesenheimer Str. 19, 67661 68169 Mannheim condition of the grease correlates with the wear condition of the gear components. A flank-critical operating point with a high transmission ratio is tested in a planetary gear. The tested operating point causes surface pressures of over 1 GPa on the tooth flank, and the narrow spaces in the planetary bearing lead to high shear stress on the grease. After the gear test, Grease A, as designated in this paper, showed complete physical degradation of the thickener, as shown in Figure 1. Table 1 shows the relevant characteristics of Grease A. Transferability to model tests With the aim of reproducing the physical thickener degradation from Figure 1 of Grease A at the model level, various tests are performed on Grease A. According to the state of the art, established laboratory and model test methods already exist. The tested methods and the optical and haptic condition of Grease A after each test are shown in Table 2. Figure 2 shows examples of the grease condition described in Table 2 after the FZG, FE8 test, and the oxidation roll tester. Grease A has become significantly discolored and softer. The thickener structure is intact, and reversible behavior of the consistency can be observed upon cooling. The tests performed in Table 2 do not cause the damage pattern of total physical thickener degradation of Grease A in the gearbox shown in Figure 1. Science and Research 51 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0019 Figure 1: Condition of Grease A after planetary gear testing: Total physical degradation from (Ochs, et al., 2024). Figure 2: Grease condition Grease A according to FZG (left), FE8 (center), oxidation roll tester (right). Name NLGI Base oil viscosity Base oil type Thickener Additive Grease A 2 100 mm 2 / s Synthetic oil Calcium sulfonate complex A1 Table 1: General characteristics of Grease A. Test Grease condition Test Grease condition FE9 Roling bearing test DIN 51821 A/ 1500/ 6000-120 Softer, oily -consistent Work stability DIN ISO 2137 Softer, oily -consistent FZG FLP-TF-0520 A/ 2,8/ 50 Softer, oily -consistent VKA Welding load DIN 51350-4 Softer, oily -consistent FE8 DIN 51819-2 C-75-80 Softer, oily -consistent Oxidation Roll Tester ASTM D 1831 Softer, oily -consistent Shell-roller test 72h/ 100°C Pw nach ASTM D 1831 Softer, oily -consistent Klein´s shear Mikro PU / mikro PW60 [0,1mm] Softer, oily -consistent Work penetration DIN ISO 2137 Softer, oily -consistent Table 2: Performed model tests and their results with Grease A. temperature during the test is kept constant at T = 100 °C. The test duration is 20 million revolutions. Adapted grease formulation and validation The modified model test is used as a method for the predevelopment of an adapted grease formulation. In order to minimize the influence of the thickener type on the functionality of the grease in the gearbox, the thickener type of Grease A is kept unchanged. A fully synthetic base oil is used, the base oil viscosity is increased to 150 cst, and the thickener structure, including the additive package, is developed further. Table 3 shows the basic characteristics of Grease B compared to Grease A. Figure 4 shows Grease A and B before and after the test on the modified VKA from Figure 3. Both greases have Science and Research 52 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0019 Modified model testing Since it is not possible to replicate the degradation process of Grease A in the gearbox using the model tests carried out in Table 2, a modified VKA (four-ball tester) was developed. A parameter study was used to specify the test setup and operating conditions that reliably replicate the degradation process of Grease A. The setup is typically used to investigate the shear stability of oils according to CEC L45-A-99 and is shown in Figure 3. The test setup based on a VKA is extended by a temperature-controlled roller bearing holder with a cover. A tapered roller bearing from the 32008 series is used as the test bearing. The test bearing is fully filled with grease and operated at a constant speed of n = 1500 min -1 . The load is 10 kN (corresponding to a pressure on the outer ring of approx. p = 1.5 GPa). The Figure 3: Modified four-ball-tester setup. Figure 4: Condition of Grease A (left) and Grease B (right) before and after testing in the modified VKA. Name NLGI Base oil viscosity Base oil type Thickener Additive Grease A 2 100 mm 2 / s Semi-synthetic oil Calcium sulfonate complex A1 Grease B 1 150 mm 2 / s Synthetic oil Calcium sulfonate complex A2 Table 3: General characteristics of Grease A and Grease B. changed not only in color but also in terms of their texture. It should be noted that the change in Grease A (left) is irreversible. It is clear to see that the thickener has been destroyed by the shear in the tapered roller bearing. In comparison, Grease B shows an expected discoloration and a normal softening, which is reversible upon cooling. To validate the suitability of the modified VKA as a new model test for assessing the grease condition in the gearbox and to validate the thickener stability of the newly developed Grease B, Grease B is tested in the same gearbox test as Grease A from (Ochs, et al., 2024). After a running time of 300 hours, the gearbox is disassembled and the grease condition and gearbox components are analyzed. Figure 5 shows the grease condition in the planetary gearbox after the test run. After the test, the grease is clearly discolored and softened but still has an oily consistency. Compared to the condition of Grease A after the same gear test in Figure 1, the thickener is intact and the softening is reversible. Discussion During gear tests according to (Ochs, et al., 2024), Grease A and B exhibit moderate operating temperatures (60 °C -70 °C). This leads to the assumption that high shear stresses are the most dominant factor reducing the remaining service life of the greases and that the chemical degradation mechanisms can be considered negligible. Furthermore, the temperatures and temperature curves do not indicate any abnormal wear in the narrow spaces between the planetary bearings, so that the flow behavior of Grease B (and the degraded Grease A) appears to be sufficient for the conditions tested. Due to the unremarkable wear condition of the bearing components after the gear tests with Grease A, it can be assumed that the effect of thickener degradation is compensated by the associated high flow behavior of the resulting grease mixture and is not yet critical for the operating time of 300 hours. However, during further operation, it can be assumed that the inhomogeneous grease mixture will have lost a significant amount of its thermal conductivity and will cause increased temperature and friction. On the other hand, the degraded thickener residues may increase the oxidation reactions in the base oil/ grease mixture, so that the improved Grease B with an intact thickener structure can be attributed a longer residual service life. In order to simulate the effect of total physical thickener degradation of Grease A in the gearbox at the model level, it becomes apparent that both the mechanical load and the available grease reservoir and its oil supply to the tribological contact are relevant. Modification of the established model tests with regard to test setup, test equipment, and operating conditions is essential for effect replication. The high shear stresses and the modified oil supply from the grease reservoir to the tribological contacts in the modified VKA cause the thickener structure of Grease A to fail, as in the gearbox test, but not the improved Grease B. Conclusion It is shown that the physical degradation of a grease in a real gearbox application could not previously be reproduced in established model tests. Simulating the effect at the model level is essential in order to optimize the thickener structure for mechanical stability during the early stages of grease development. The model test developed in this work on a modified VKA allows the total physical thickener degradation to be isolated and reproduced for the time being. A new grease with an adapted thickener system proves to be mechanically stable on the modified VKA. The stability of the new thickener system and thus the suitability of the modified model test for early development was validated in a real gearbox application. The modified VKA is proposed as a new method for the specific investigation of the mechanical stability of the thickener structure of grease. References (1) Kuhn Erik Zur Tribologie der Schmierfette [Buch]. - Renningen : expert verlag, 2017. - Bde. Eine energetische Betrachtungsweise des Reibungs- und Verschleißprozesses. (2) Meijer Robert Jan, Osara Jude A. und Lugt Piet M. On the Required Energy to Break Down the Thickener Structure of Lubrication Greases [Journal] / / Tribology Transactions. - [s.l.] : Taylor & Francis, 2023. (3) Ochs Georg [et al.] Experimentelle Untersuchung der Übertragbarkeit von Fettkenngrößen auf fettgeschmierte Getriebe [Journal] / / Tribologie und Schmierungstechnik. - [s.l.] : Narr Francke Attempto Verlag GmbH + Co. KG, 2024. (4) Rezasoltani Asghar und Khonsari M.M On Monitoring Physical and Chemical Degradation and Life Estimation Models for Lubricating Greases [Journal] / / Lubricants. - 2016. Science and Research 53 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0019 Figure 5: Condition of Grease B after planetary gear testing under operating conditions (Ochs, et al., 2024). deposition techniques, which include Physical Vapour Deposition (PVD), chemical vapor deposition, and ion beam assisted deposition, can produce more homogeneous surfaces, but surface preparation and costs are higher [2, 3]. Molybdenum and its oxides have been studied as a good candidate for solid lubricant coating technology because of their higher hardness and relatively low wear coefficient when compared to uncoated bearing surfaces [4]. In recent studies by our research group, the wear and hardness of molybdenum-based coatings applied with PVD were evaluated from nanoto macroscopic scales [5, 6, 7], and it was shown that molybdenum can work in dry bearing surfaces under ambient conditions [8]. However, PVD presents limitations in coating thickness (approx. 2 µm) [3]. As an alternative, APS can provide a thicker coating (approx. 50 µm to 500 µm), leading to Science and Research 54 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0020 Introduction Solid lubricant coatings can be a valuable alternative to fluid lubricants when these can’t be used, particularly in extreme conditions like high vacuum, extreme temperatures, or contamination-sensitive environments [1]. Solid lubricant coatings face several challenges that must be addressed to ensure their effective functionality and performance. These include high friction coefficients and wear compared to elasto-hydrodynamic lubrication [1]. Additionally, once worn, solid lubricant coatings cannot usually regenerate, leading to the contacting materials being different from the initial ones and possibly leading to higher friction, wear, and consequently failure. To apply such coatings, different production processes can be used: thermal spraying methods and deposition methods [2]. Thermal spraying methods include, for instance, Atmospheric Plasma Spraying (APS) and flame spraying. Compared to deposition methods, thermal spraying methods are cost-effective, capable of delivering high deposition rates, and are suitable for producing thick coatings, but have the disadvantage of producing a more porous microstructure [2]. On the other hand, Molybdenum APS-Coatings: A Self-regenerative Solution for Wear Resistance in Dry-lubrication Applications Ricardo Crespo Martins, Dennis Konopka, Mareike Dukat, Florian Pape, Martin Nicolaus, Kai Möhwald, Gerhard Poll, Max Marian* Presented at GfT Conference 2025 Molybdenum coatings and their oxides are a promising technology to protect bearing surfaces from wear due to their good wear-resistance and their capability to regenerate by forming new protective and lubricious oxide layers. However, until recently, their applicability against higher hardness materials has not yet been studied. This study aims to explore Atmospheric Plasma Spraying (APS) to produce a molybdenum layer on bearing steel and compare the tribological behavior of coated and uncoated surfaces using a reciprocating pin-on-plate apparatus under dry-lubrication. The counter-body material was chosen to be alumina, given its inertness and high hardness. The tests were conducted at various distances and loads, comparing the friction and wear volumes of coated and uncoated specimens. Keywords Coatings, Wear-Resistant, Thermal Spraying, Solid Lubrication, DFG SPP 2074. Abstract * Ricardo Crespo Martins (corresponding author) 1 Dennis Konopka 1 Mareike Dukat 2 Dr. Florian Pape 1 Dr. Martin Nicolaus 2 Prof. Kai Moehwald 2 Prof. Gerhard Poll 1 Prof. Max Marian 1,3 1 Institute of Machine Design and Tribology, Leibniz Universität Hannover, An der Universität 1, 30823 Garbsen, Germany. 2 Institute of Materials Science, Leibniz Universität Hannover, An der Universität 2, 30823 Garbsen, Germany. 3 Department of Mechanical and Metallurgical Engineering, School of Engineering, Pontificia Universidad Católica de Chile, Vicuña Mackenna 4860, 6904411 Macul, Región Metropolitana, Chile. longer dry-running times and lower maintenance costs. It also has the advantage of being able to regenerate due to the capability of molybdenum to oxidize with the ambient oxygen, when previous oxide layers wear out [9]. However, APS has shown lower hardness and higher porosity compared to PVD [3]. In the present work, the wear and friction behaviors of bearing steel and APS-sprayed molybdenum are compared under reciprocating conditions against a counter body made of alumina, at a wider range of loads and distances than in the previous study [3]. With this research, we aim to understand whether APS-applied molybdenum still outperforms bearing steel when a hard material is slid on its surface. Methodology The study was conducted using coated (hereafter, APS) and uncoated (hereafter, REF) specimens. The uncoated specimen and the substrate of the coated specimen were made of hardened 100Cr6 (AISI 52100) steel discs with a diameter of 40 mm and a height of 10 mm. To apply the APS coating, one of the flat surfaces of the discs was roughened by abrasive blasting with corundum of a grain size of 250 to 355 µm as blasting medium. The surface was then cleaned with isopropanol to remove contaminants. The coating was applied with atmospheric plasma spraying using a Delta-torch (GTV Verschleißschutz GmbH, Germany). The coating material was molybdenum powder (Mo with 99 % purity, powder particle size between 45 and 90 µm; GTV Verschleißschutz GmbH, Germany). The powder was melted using a plasma jet made of an ionized mixture of argon and hydrogen generated by an electric arc. The molten molybdenum was then projected to the substrate surface, solidifying and adhering to it. The APS process parameters are presented in Table 1. After the coating process, the surface was ground to a final thickness of 50 µm using silicon carbide as abrasive slurry. This thickness ensures that the maximum stress due to the non-conforming contact happens below the coating [10]. Before being tribologically stressed, the samples were cleaned using acetone and isopropanol. To evaluate the coated and uncoated systems, a ballon-plate apparatus (Milli-Tribometer, TRIBOtechnic, France) was used. With this tribometer, a loaded sphere (counter-body) was slid in a reciprocating motion against the specimen, and the friction force was measured. The conditions used in the tribological tests are given in Table 2. Both the load and sliding distance were varied, totaling nine sets of conditions for each of the specimen materials. For each test condition, three repetitions were carried out. This number of repetitions may only be of statistical relevance if the behavior (wear volume and friction coefficient) is similar between each repetition and the difference between conditions is considerable. The counter-body had a diameter of 6 mm, and was made of alumina (99.8 % Al 2 O 3 , G10; TK Linear GmbH, Germany). This material was chosen due to the particular high hardness, high wear resistance, and inertness. In previous studies [3], steel counter-bodies were used and the wear occurred on these. Although alumina is less common in rolling element bearings, it provides a mean to evaluate the self-regenerative properties in harsher conditions compared to the conditions provided by steel counter-bodies. Given the material properties and geometry of the contacting bodies, the initial Hertzian contact pressures resulted in 764, 963, and 1102 MPa, which respectively correspond to 1, 2, and 3 N of applied load. After being tribologically stressed, the samples were analyzed under a confocal laser scanning microscope (LSM; VKX-100, Keyence, Japan). This microscope enables the measurement of the surface profile and, consequently, the determination of the wear volume. To compare the wear behavior of the coated and uncoated samples, the surface profile of each wear scar was analyzed at three distinct locations. The wear area of each section was then determined and averaged per scar and test condition (totaling nine profile measurements per condition). This average wear area was then multiplied by the length of the scar (5 mm) to determine the wear volume. Results and Discussion The results of the friction coefficient corresponding to each distance at which the test was stopped are presented in Figure 1. As mentioned in the methodology, three distances were tested: 5, 50, and 100 meters, each one cor- Science and Research 55 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0020 1 3 62$27,#,$* 82.',* \#0*-)$*&I'))0,-& OWW&B& \#0*-)$*&F"30)& SW&; Z& B).",&2#"3&)+-0&& OW&cd: $,& U8%)".0,&2#"3&)+-0& >W&cd: $,& F"3%0)&200%&)+-0& I+@&EQ<QW&.d: $,& \9-0),+#& *""#$,.& "2& -40& / +: 7#0/ & B*-$10& L$/ -+,*0& (0-300,& ,"XX#0&+,%&/ '(/ -)+-0& >OW&: : & Table 1: Atmospheric Plasma Spraying parameters. 62$27,#,$ & 82.', & c"+%&& >6&E6&M&f& N+9@&/ #$%$,.&/ 700%& T&: : d/ & ! -)"; 0& Q&: : & ! #$%$,.&%$/ -+,*0& Q6&QW6>WW&: & I"',-0)<("%8&%$+: 0-0)& R&: : & Table 2: Test conditions used in the pin-on-plate tribometer. observed to increase. While the APS specimens demonstrated to have a higher friction coefficient at higher loads, the highest friction coefficient for REF was measured at the middle load (2 N), and the lowest was recorded at the highest load (3 N). In the bottom figure (Figure 1c), the same distinct behavior of increasing and decreasing friction values for REF and APS, respectively, is noticeable. However, unlike the tests stopped at 50 meters, the friction coefficient of REF was measured as the highest with the highest load, and vice versa. As for the APS, the friction coefficient values were shown to converge to a value of around 0.65. It was found that the longer the test, the lower the average standard deviation. While in the tests lasting 5 m the standard deviation was high, for the 100 m tests, the highest average standard deviation was recorded as 0.07 for REF at a load of 3 N. After each test, the surface profile of each wear scar was analyzed. The results of the average wear volume per test are shown in Table 3. Whenever it was possible to measure a wear area, the values are given, otherwise, a zero value is presented. The data indicate that wear was always measured for REF, regardless of the distance or Science and Research 56 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0020 responding respectively to the figures a), b), and c). Each line corresponds to the average friction coefficient per test condition. The running-in behavior of the two specimens at three different loads is depicted in Figure 1a). It is shown that the friction coefficient of both specimens starts at a low value, around 0.2, and it is shown to increase steadily. While REF was observed to take more time to reach this behavior, less time was required by the APS specimen, and after around 0.5 meters of the test, its friction coefficient is already increasing less abruptly. In these first meters, it is observed that the friction coefficient of REF is already higher at higher loads (2 and 3 N), but lower at the lowest tested load (1 N). It should be noted that the running-in behavior had a large standard deviation (from the six lines of the graph, the highest average standard deviation across the whole test was 0.13 for REF at a load of 3 N), making the comparison between each condition less reliable. In a second phase of the test (tests that lasted 50 meters), it was observed that the friction coefficients of the specimens exhibited different behaviors: while APS tests were noted to decrease with the distance, REF tests were ! ! "# $"# %"# Figure 1: Friction coefficient evolution with distance of a 6 mm diameter reciprocating ball-on-plate tests at 8 mm/ s. Each line corresponds to the mean of three repetitions. load, except at the lowest load and distance. However, for APS, at lower loads, the wear was found to be very low or negligible at any distance. Except for the tests that lasted 50 meters with a load of 2 N, APS has always shown to have a lower wear volume than REF. Exemplary images of the APS and REF surfaces taken using the LSM are given in Figure 2a) and 2b), respectively. After the tests, the counter-bodies were analyzed under a microscope and it could be concluded that no wear occurred on these, as it is possible to see in Figure 2c) and 2d). However, for the APS tests, a layer on the surface of the sphere was noticeable, possibly molybdenum and its oxides. As the tests had three different distances, it was possible to evaluate the friction and wear at the end of each distance. After a running-in phase during the first meters of test, the contacting materials were the ones expected (steel and its oxides with alumina; molybdenum and its oxides with alumina). Given the differences in hardness between the molybdenum (3.8 GPa) and steel surfaces (typically 8 GPa), and the counter-body (typically 15-16 GPa), it would be expected that the wear volume of the molybdenum would be higher. However, even at lower distances and a low amount of repetitions, it was observed that this was not the case, and the coated specimens presented lower wear. It was also evident that at low loads (1 N), regardless of the sliding distance, wear was not observed on the coated specimen, a phenomenon that was only seen at low distances and loads with the uncoated specimen. This is of particular relevance, given the high hardness of the counter-body. This phenomenon of the molybdenum coating points to a high dependence Science and Research 57 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0020 Figure 2: Surfaces of the specimens (a and b) and counter-bodies (c and d) after a 100 m test with a load of 1 N at 8 mm/ s. The direction of sliding is shown by the red arrow. Table 3: Average wear volume of the specimen in x10 -6 mm 3 . d) Surface of the counter-body after the test against the uncoated specimen (REF). c) Surface of the counter-body after the test against the coated specimen (APS). a) Wear scar of the coated specimen (APS). b) Wear scar of the uncoated specimen (REF). Mo-based solid lubricant coatings deposited by APS as a PVD alternative: Mechanical and tribological performance,” Tribology Transactions, just-accepted, pp. 1-18, 2025, doi: 10.1080/ 10402004.2025.2519334 [4] M. Marian, D. Berman, A. Rota, R. L. Jackson, and A. Rosenkranz, “Layered 2d nanomaterials to tailor friction and wear in machine elements - a review,” Advanced Materials Interfaces, vol. 9, no. 3, p. 2101622, 2022, doi: 10.1002/ admi.202101622 [5] B.-A. Behrens, G. Poll, K. Möhwald, S. Schöler, F. Pape, D. Konopka, K. Brunotte, H. Wester, S. Richter, and N. Heimes, “Characterization and modeling of nano wear for molybdenum-based lubrication layer systems,” Nanomaterials, vol. 11, no. 6, p. 1363, 2021, doi: 10.3390/ nano11061363 [6] N. Heimes, F. Pape, G. Poll, D. Konopka, S. Schöler, K. Möhwald, and B.-A. Behrens, “Characterisation of selfregenerative dry lubricated layers on Mo-basis by nano mechanical testing,” in Production at the leading edge of technology: Proceedings of the 9 th Congress of the German Academic Association for Production Technology (WGP), September 30th-October 2nd, Hamburg 2019. Springer, 2019, pp. 139-148, doi: 10.1007/ 978-3-662-60417-5_14 [7] S. Schöler, M. Schmieding, N. Heimes, F. Pape, B.-A. Behrens, G. Poll, and K. Möhwald, “Characterization of molybdenum based coatings on 100Cr6 bearing steel surfaces,” Tribology Online, vol. 15, no. 3, pp. 181-185, 2020, doi: 10.2474/ trol.15.181 [8] D. Konopka, F. Pape, N. Heimes, B.-A. Behrens, K. Möhwald, and G. Poll, “Functionality investigations of drylubricated molybdenum trioxide cylindrical roller thrust bearings,” Coatings, vol. 12, no. 5, p. 591, 2022, doi: 10.3390/ coatings12050591 [9] Maier, Hans, Kai Möhwald, Gerhard Poll, and Florian Pape. “Dry Lubrication of Rolling Bearings.” DE Patent 102023101922A1, filed August 2024. [10] T. Coors, F. Pape, and G. Poll, “Comparing the influence of residual stresses in bearing fatigue life at line and point contact,” Residual Stresses 2018: ECRS-10, vol. 6, p. 215, 2018, doi: 10.21741/ 9781945291890-34 [11] B. Hwang, J. Ahn, and S. Lee, “Effects of blending elements on wear resistance of plasma-sprayed molybdenum blend coatings used for automotive synchronizer rings,” Surface and Coatings Technology, vol. 194, no. 2-3, pp. 256-264, 2005, doi: 10.1016/ j.surfcoat.2004.07.072 [12] S. Manjunatha and S. Basavarajappa, “Effect of powder particle size on wear resistance of plasma sprayed molybdenum coating,” Proceedings of the Institution of Mechanical Engineers, Part J: Journal of Engineering Tribology, vol. 228, no. 7, pp. 789-796, 2014, doi: 10.1177/ 1350650114531744 [13] I. K. Piltaver, I. J. Badovinac, R. Peter, I. Saric, and M. Petravic, “Modification of molybdenum surface by lowenergy oxygen implantation at room temperature,” Applied surface science, vol. 425, pp. 416-422, 2017, doi: 10.1016/ j.apsusc.2017.07.029 [14] J. Khedkar, A. Khanna, and K. Gupt, “Tribological behaviour of plasma and laser coated steels,” Wear, vol. 205, no. 1-2, pp. 220-227, 1997, doi: 10.1016/ S0043-1648(96)07291-2 Science and Research 58 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0020 of the wear on the load, as confirmed by [11, 12]. The lower wear of the APS-coated specimen is justified by the formation of a tribofilm composed of molybdenum oxides. These oxides are formed in the presence of ambient oxygen near the wear scar and the higher temperature generated by the friction energy, which would allow for better conditions for the oxides to form [13]. These oxides, when formed, can serve as a lubricant due to their high hardness and 2D layered structure that provides low-shear films and thus lower friction [14]. It is postulated that when these oxides are removed from the surface, new ones are formed, self-regenerating the surface with the lower friction material. Conclusions and Outlook Molybdenum coatings applied using APS show good wear-resistant properties and a lower friction coefficient than the uncoated bearing steel, when a hard and inert counter-body is slid on their surface. This coating is a good candidate as a technology that can be used in highly loaded contacts and to enhance the tribological behavior of rolling element bearings under dry lubrication, even when hard counter-bodies are used. Although it is more common for steel spheres to be used, these conditions can be seen in hybrid rolling element bearings when the fluid lubrication system fails and no lubricant is available to separate the surfaces, as well as in conditions where fluid lubrication is not suitable due to extreme environmental conditions. Given that the amount of molybdenum necessary to apply the coating is small and the time to prepare and apply the APS coating is also short when compared with other coating production technologies (PVD, for instance), this coating should be evaluated as a technology to be used in bearing raceways. In future work, the behavior of this coating technology shall be evaluated under various ambient conditions to better understand the dependence of oxide formation with temperature, humidity, and pressure. Tests with rolling element bearings featuring molybdenum coated surfaces will help assess the viability of the coating and the method of its application. References [1] E. Omrani, P. K. Rohatgi, and P. L. Menezes, Tribology and applications of self-lubricating materials. CRC Press, 2017. [2] V. Kumar and B. Kandasubramanian, “Processing and design methodologies for advanced and novel thermal barrier coatings for engineering applications,” Particuology, vol. 27, pp. 1-28, 2016, doi: 10.1016/ j.partic.2016.01.007 [3] D. Konopka, R. Crespo Martins, M. Dukat, F. Pape, K. Möhwald, G. Poll, and M. Marian, “Self-regenerative Science and Research 59 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0021 2 Materials and Methods In order to systematically investigate the friction and wear properties of all material-oil-combinations tribological tests were conducted using a pin-on-disc tribometer. Coated steel pins manufactured from tool steel 1.2379 were brought into contact with stainless steel sheets (1.4404) to simulate realistic conditions. All combinations were tested under identical conditions. The following section outlines the tribological test setup, the coatings applied, and the lubricants and evaluation parameters considered. 2.1 Tribological Testing The pin-on-disc tests were carried out on a BRUKER TriboLab universal tribometer. Evaluated parameters in- 1 Introduction Tool and workpiece materials are subject to extraordinary load in the field of sheet metal forming. Lubricants deliver friction reduction, wear protection and overall process stability. Conventional lubricants, however, are increasingly criticised due to their fossile origin and low decomposability. Bio-based lubricants originating, especially coming from residual material, are a sustainable alternative. Oil from Hermetia Illucens (the black soldier fly) is a byproduct of waste treatment ad can be transformed into technical ester. In addition, bio-based additives from various agricultural or plant byproducts can also be utilized. These components must prove equal lubrication capabilities compared to conventional lubricants in order to show eligibility for industrial production environments. Furthermore, there is a high interest into the interaction with plasma vapor deposited and diffusion treated surfaces. This work investigates and compares the applicability of lubricants based on oil of the black soldier fly Hermetia Illucens and plants and the effect of various added conventional and bio-based additives. Consideration is laid on the interaction of different surface modifications of tool steel 1.2379 with fine steel 1.4404. Comparing studies inspect the eligibility of Hermetia based oil as well as the interplay of various plant-based and synthetic additives with different tool surface modification. Aim of this work is to describe the influence of these surfaces on the specific additives. The different plasma vacuum deposited coatings and modifications differ in chemical bonding, surface energy and reactivity. These properties highly influence the additive effects and suitable adjustment may optimize the lubrication effect. Influence of tool coatings on the tribological effectiveness of biolubricants for sheet metal forming Hoang Viet Le, Kai Weigel* Presented at GfT Conference 2025 Environmentally friendly lubricants are essential for sustainable sheet metal forming. Their performance depends on interactions between additives and tool surfaces. This study evaluates bio-based TMP-esters, conventional and bio-based additives, and plasma vacuum deposited coatings. Lubricants and coatings were tested in a pin-on-disk setup at different temperatures in oscillating and rotating mode. Results indicate that bio-based lubricants from Hermetia illucens and plants are suitable for sheet metal forming. Biobased additives further show promising potential. The investigated coatings demonstrated excellent tribological properties, both independently and in combination with bio-based lubricants, supporting their use in sustainable forming processes. Keywords sheet metal forming, lubricant additives, tool coatings, friction and wear, bio-based lubricants, pin-ondisk test, additive-surface interaction, resource-efficient tribsystems Abstract * Hoang Viet Le, M.Sc. Dipl.-Ing. Kai Weigel Fraunhofer-Institute for Surface Engineering and Thin Films IST Riedenkamp 2, 38108 Braunschweig, Germany clude the coefficient of friction, friction track width on the sheet surface, stick-slip behavior, and wear on the pin surface. The influence of lubricants, additives, and tool coatings was analysed under conditions closely resembling industrial practice. The test setup and parameters were derived from preliminary experiments that showed good agreement with real forming processes (see Table 1). 2.2 Tool coatings and Counter body An austenitic stainless steel of type 1.4404 was used as the counter body. This material is widely applied in forming technology and commonly used in the automotive industry. It is characterised by high strength and good corrosion resistance. Typical tribological challenges, however, include a pronounced tendency for adhesion and strain hardening during forming, which can lead to increased tool wear. Science and Research 60 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0021 Parameters Unit Test values Normal force [N] 25 Temperature [°C] 30 / 110 Motion type Oscillating, rotating Stroke length [mm] 20 Stroke frequency [Hz] 1,25 Duration per load step [s] 600 Pin Ø 10 / tool steel 1.2379 58 HRC Sheet 50x40 x1 mm, Nirosta (1.4404) Table 1: Test parameters of the pin-on-disc tribometer tests performed Classification Lubricant Additives Description Reference KTL N16 Unknown Mineral oil-based fully formulated lubricant Bio-based base oils TMP-ester 1 None Based on Hermetia illucens TMP-ester 2 None Plant-based TMP-ester 3 None Bio-based base oils with 5% additive TMP-ester EP additive 1 EP additives with different sulfur contents TMP-ester EP additive 2 TMP-ester EP additive 2 TMP-ester Additive package 1 Commercial additive packages TMP-ester Additive package 2 TMP-ester Bio additive 1 Bio-based additives TMP-ester Bio additive 2 TMP-ester Bio additive 2 Bio-based base oils with 10 % additive TMP-ester EP additive 1 EP additives with different sulfur contents TMP-ester EP additive 2 TMP-ester Bio additive 1 Bio-based additive Table 3: Investigated lubricants, additives and description Type of coatings/ surface treatment Selected variants Dominant attachment type Diamond-like carbon (DLC) coatings a-C: H Covalent, weakly polar a-C: H: Si Covalent, weakly polar a-C: H: W Covalent, moderately polar Carbide and nitride coatings WC Covalent, moderately polar CrN Covalent, strongly polar Oxide coatings ZrO 2 Weakly ionic Diffusion treatment with nitrogen Fe X N Y Covalent, strongly polar Table 2: Investigated Coatings and Surface Treatments Science and Research 61 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0021 The test pins, representing the tool material, were made of tool steel 1.2379 (X153CrMoV12) with a hardness of 60-63 HRC. Prior to coating deposition, the pins underwent a plasma nitriding process followed by dry electropolishing in order to enhance coating adhesion and optimise tribological performance. A series of hard coatings with different chemical bonding characteristics was applied. Table 2 provides an overview of the coating systems and their predominant bonding types. 2.3 Lubricants and Additives A fully formulated mineral oil (Raziol KTL N16) was used as the reference lubricant. The exact composition of this oil is not disclosed by the manufacturer. KTL N16 is well established in sheet metal forming and therefore serves as an industry relevant lubricant. In addition, Hermetiaand plant-based TMP-esters of different origins were used as base oils for additive formulation. Additives were blended into the base oils at concentrations of 5 wt.% and 10 wt.% (see Table 3). The following types of additives were investigated: • Extreme pressure additives • Commercial additive packages • Bio-based additives derived from plant residues 3 Results To evaluate the tribological behavior, both friction and wear parameters were analysed. Friction was assessed based on the average coefficient of friction. The wear analysis included the friction track width on the sheet surface, the occurrence of stick-slip effects in the form of chatter marks, as well as the size and extent of the wear zone on the spherical pin surface representing the forming tool. The results are presented below. The comparison (oscillation tests) with the fully formulated reference lubricant KTL N16 shows that bio-based base oils achieve comparable friction levels at 30 °C. At higher temperatures, particularly at 110 °C, the coefficient of friction of the reference decreases due to additive activation, whereas the unadditivated bio-based base oils maintain higher friction levels (see Fig. 1). This results in an increasing deviation from the reference, highlighting the necessity of targeted additive formulation. Only the combination of a bio-based base oil with the EP additive 1 (TMP-EP) achieved friction values on the same level as the reference lubricant. A higher additive concentration did not yield further improvement. All other bio-based additives exhibited lower tribological performance across all temperature levels. The choice of tool coating also had a pronounced effect on the tribological behaviour (see Fig. 1, right). Carbon-based coatings, in particular a-C: H: Si, led to a significant reduction in friction regardless of the lubricant applied and thus confirmed the effectiveness of these coatings. Other coatings achieved only selective improvements in separate tests, depending on the additive type. In addition to the oscillating tests, initial rotating tribometer experiments were performed. In these tests, the bio additive 1 in the bio-based base oil achieved the best friction performance of all lubricants tested, including the reference. This indicates a high tribological potential of plant-based additives under rotating load conditions. Figure 2 shows the friction track widths on the sheet surface at 30 °C, 70 °C and 110 °C after completing all load stages with both uncoated and coated pins. The results for friction track widths largely follow the same trends as the friction values. Notably, differences Figure 1: Average coefficient of friction for various lubricant-surface systems tely avoided with a-C: H. This highlights the effective role of carbon-based coatings in preventing stick-slip effects. The size and extent of the wear zone on the spherical pin surface provide an indication of the expected tool wear. As a result of the groove-shaped friction track on the sheet surface, an oval contact zone develops on the pin surface, the width of which generally corresponds to the friction track width. The wear zone predominantly shows abrasive marks and is often superimposed by a linear zone with adhesion of sheet material. Under effective lubrication, both areas are only weakly pronounced. While the uncoated variant exhibits a clearly developed wear zone with visible material adhesions, no adhesions are observed with a-C: H. 4 Discussion The key findings regarding friction behaviour, wear mechanisms and the interaction between lubricants and tool coatings are summarised below: Science and Research 62 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0021 between the fully formulated reference lubricant and the bio-based base oils are already evident at 30 °C, whereas for the friction values these differences only became pronounced at elevated temperatures. Apart from this, the friction track widths confirm the same tendencies as the friction values: only the combination of a bio-based base oil with an EP additive reached levels comparable to the reference, while all other bio-based additives showed lower performance across all temperature stages. The positive effect of the carbon-based coating a-C: H: Si was also clearly confirmed by the reduced friction track widths. The supplementary rotating tests revealed a similar overall picture, although the differences were less pronounced compared to the friction values. Stick-slip effects occur when the lubricant provides insufficient separation between the contact partners and manifest in the form of chatter marks. These effects can be influenced by suitable coatings. Figure 3 illustrates exemplary friction tracks of one lubricant in combination with an uncoated pin and with an a-C: H coating. While pronounced chatter marks appear in all temperature stages with the uncoated variant, they are comple- Figure 2: Friction track widths on the sheet metal surface for different lubricant-surface systems Figure 3: Scratch marks and pen wear - uncoated vs. a-C: H • The addition of EP1 led to a significant reduction in the coefficient of friction and the friction track width. Without additive, the tribological parameters were clearly higher than those of the fully formulated reference lubricant (KTL N16). • Carbon-based coatings such as a-C: H and a-C: H: Si showed a significant reduction in friction and wear, independent of the lubricant applied. These layers apparently act as friction-reducing boundary films, even without specific chemical interaction with the additive. • The comparison with a-C: H: W and WC revealed that a higher tungsten content in the coating can increase the tendency for chatter mark formation and track wear, particularly at elevated temperatures. • Stick-slip effects occurred primarily on uncoated pins. Their occurrence was suppressed by suitable coatings, indicating more stable interfacial friction. • The wear zones on the pin surfaces demonstrated that the combination of carbon-based coatings and additives not only reduces friction but also leads to lower adhesion and abrasive attack. • The investigated bio-based additives did not show a consistent tribological effect in the initial tests. Further studies are required to assess their reactivity and protective performance under different conditions. 5. Conclusions The results demonstrate that bio-based lubricating greases derived from Hermetia illucens can provide a lubrication performance comparable to conventional lubricants. Bio-based lubricants also allow for highly effective additive formulation. Amorphous carbon coatings reduce friction and tool wear regardless of the lubricant type. Future investigations should include additional materials to evaluate the transferability to different forming scenarios. Furthermore, the targeted optimisation of additive combinations as well as tests under more applicationoriented conditions - such as elevated temperatures, continuous sliding motion, or forming trials - will be essential. The described work was accomplished in coorperation of the Fraunhofer Institutes IST, IVV, IWU, IME and UMSICHT. Reference [1] Murrenhoff, H.: Umweltverträgliche Tribosysteme - Die Vision einer umweltfreundlichen Werkzeugmaschine. Berlin: Springer, 2010. [2] Fachagentur Nachwachsende Rohstoffe; Peterek, G. (Mitarb.): Bioschmierstoff-Kongress: 12.-13. November 2014, Hagen. Gülzow-Prüzen: Fachagentur Nachwachsende Rohstoffe e.V., 2015 (Gülzower Fachgespräche, Band 50). [3] Wang, Y.-S.; Shelomi, M.: Review of Black Soldier Fly (Hermetia illucens) as Animal Feed and Human Food. Foods (Basel, Switzerland), 6 (2017) 10. [4] Holweger, W.: Wechselwirkung von Additiven mit Metalloberflächen. 2., überarb. u. erw. Auflage. Renningen: expert-Verlag, 2022 (Tribologie - Schmierung, Reibung, Verschleiß). [5] Maßmann, T. Ch.: Wirkmechanismen additivierter Schmierstoffe in der Kaltumformung. 1. Auflage. Aachen: Shaker, 2007 (Berichte aus der Produktionstechnik, 2007/ 18) [6] Le, H. V.: Einfluss von Werkzeugbeschichtungen auf die tribologische Wirksamkeit von Schmierstoffadditiven für die Blechumformung. Masterarbeit, Technische Universität Braunschweig, 2025. Science and Research 63 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0021 Science and Research 64 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0022 Introduction In innovative industrial applications, especially in the automotive and drive technology sectors, the demand for sealing systems is continuously increasing. High circumferential speeds are necessary to meet new performance requirements, which in turn leads to significant thermal stress within the tribological system - the radial shaft sealing system (Figure 1). During dynamic operation, friction between the system components - the radial shaft seal (RSS), the shaft surface, and the lubricant - converts the mechanical energy generated by the contact partners into heat, with the highest temperature occurring directly within the contact zone (sealing gap). Increased contact temperatures can pose a significant thermal challenge for both the lubricant and the sealing material. In sealing systems that are not optimally designed, high contact temperatures can lead to lubricant degradation - such as the formation of deposits at the sealing edge - and damage of the sealing material, including hardening of the elastomer (Figure 4). The presence of deposits, along with a hardened sealing edge, can significantly influence the mechanical properties of the elastomer. This, in turn, can disturb the reverse pumping effect of the RSS (Figure 2), which is essential for a well performing sealing system. An impaired reverse pumping Optimization of Radial-Shaft-Sealing systems for the use at high circumferential speeds Adrian Heinl, Erich Prem, Christian Wilbs, Fabian Kaiser, Daniel Frölich* Presented at GfT Conference 2025 Radial Shaft Seals (RSS) are critical components to seal technical applications and are increasingly challenged with high circumferential speeds. In many innovative applications, e.g. within the automotive industry and drive technology, high speeds are necessary to meet performance requirements. High circumferential speeds increase tribological and thermal challenges, potentially leading to reduced sealing performance and shortened system lifetime. Therefore, this paper evaluates possible modifications to optimize the RSS-system, such as adjusting the radial load and selecting suitable elastomer materials. Using model-based approaches, contact temperatures are estimated and experimentally validated to prove the effectiveness of the modifications Keywords Simmerring, radial shaft seal, lubricant, elastomer, testing, contact temperature, temperature estimation, high-speed, optimization, ExACT, Theta, Gümbel curve Abstract * M.Sc. Adrian Heinl, Erich Prem, M.Sc. Christian Wilbs, Dr.-Ing. Fabian Kaiser, Dr.-Ing. Daniel Frölich Freudenberg FST GmbH Höhnerweg 2-4, 69469 Weinheim Figure 1: Radial-Shaft-Sealing System [1] Figure 2: Reverse pumping effect [1] effect can ultimately result in premature failure of the sealing system and leakage. To effectively reduce contact temperature during operation, minimizing friction within the sealing system is essential. This can be achieved through targeted optimization strategies, such as adjusting the shaft surface topography and selecting high-performance lubricants. In the context of radial shaft seals, further potential for improvement lies in modifying the radial load, selecting suitable elastomer materials and optimizing the seal design. Modeling Approaches for Contact Temperature While dynamic testing is the most direct method for evaluating the effectiveness of optimization measures, it is often resource-intensive and time-consuming. The contact temperature itself is not easily measurable. It requires, e.g. a thermocouple to be installed in the shaft, some way of transferring the data to the stationary measurement system and careful adjustment as the sealing contact is typically less than 0.2 mm wide. Furthermore, the shaft surface needs to be re-ground after every test due to the wear on the shaft, which can only be done a limited number of times per shaft. All of this makes it impractical to measure the contact temperature for larger studies. To enable a more efficient preliminary assessment of potential optimization solutions, model-based approaches for estimating contact and oil sump temperatures can be used. These methods allow for an early-stage evaluation of the fundamental suitability of various seal modifications without the need for dynamic testing. To estimate the contact temperature in sealing systems, two methods are used: ExACT Equation The ExACT equation (Extended Approximation of the Contact Temperature) was developed at the Institute for Machine Elements (IMA), as a generalized method for estimating the contact temperature of the RSS systems [2]. This equation describes a linear relationship between the temperature increase in the sealing gap ΔT - defined as the difference between the contact temperature T contact and the oil sump temperature T oil sump - and the frictional power P R referenced to the shaft diameter d. The proportionality constant R 0 represents the thermal resistance of the system and characterizes its ability to dissipate heat to surrounding components such as the shaft and oil sump. Since the oil sump heats up during dynamic testing due to the frictional heat generated - and this affects the contact temperature - the oil sump temperature is also estimated. The estimation is performed iteratively to account for the thermal influence over time, continuing until a stable thermal condition is reached. To improve the accuracy of this estimation, Freudenberg developed a simplified heat transfer model specifically tailored to the oil chamber of the RSS test rig. Based on natural convection, the model describes heat transfer from the oil sump through the chamber housing to the surrounding environment. A more precise determination of the oil sump temperature enables a more reliable estimation of the contact temperature. Thermal Network The thermal network is based on the simulation tool “Theta”, developed as part of FVA project 574 IV [3] at the Institute for Machine Elements and Gear Technology (MEGT). This tool enables the digital replication of the RSS test rig oil chambers. By defining thermal nodes and specifying the thermal conductivities of individual components, it is possible - given the input of ∆ T = T contact − T oil sump = R 0 ∙ P R π d Science and Research 65 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0022 Figure 3: FKM seal after dynamic testing, in very good condition Figure 4: FKM seal after dynamic testing, with deposits, hardening and discoloration Method 1: Gümbel Curve Based Estimation The first method calculates frictional power using the coefficient of friction, radial load of the RSS, shaft speed, and shaft diameter. The coefficient of friction is derived from the Gümbel curve, which describes the relationship between the friction coefficient and the Gümbel number G. The latter characterizes the lubrication condition of an RSS system and is determined by the dynamic viscosity η of the lubricant, the angular velocity of the shaft ω and the contact pressure pm by the seal on the shaft [2]. To establish this relationship, friction torque measurements were conducted on RSS test rigs at Freudenberg, as shown in Figure 5. Science and Research 66 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0022 the heat source - to estimate the contact temperature as well as the oil sump temperature through heat conduction. Experimental Determination of Friction Power and the Gümbel Curve The estimation of the contact temperature is based on the modeling approaches described in the previous section. Both approaches require the input in the form of a heat source generated by the frictional loss of the sealing system. Two established methods were used to determine this frictional power. Figure 5: Cross section of the RSS friction torque test rig [1] Figure 6: Test procedure for the friction torque measurement Figure 7: Gümbel curve - Relationship between the Gümbel number (G) and the friction coefficient (μ) These measurements served as the basis for subsequent contact temperature calculations. The test series included two FKM materials - 75 FKM 585 and 75 FKM 170055 - and two different RSS geometries: the standard BAUM seal design and the Premium Sine Seal (PSS). All tests were performed using plunge-ground shafts and mineral oil (VG 220). The rotational speed range covered both low speed (0.35 m/ s) and high-speed conditions (6.3 to 18.8 m/ s), as shown in Figure 6. Figure 7 shows the relationship between the Gümbel number G and the friction coefficient μ, based on dynamic friction torque measurements under varying operating conditions. A quadratic approximation function, as proposed by Feldmeth [2], was used to formulate a regression equation that serves as the basis for determining friction torque. Method 2: Friction Torque Based Estimation The second method bypasses the use of the Gümbel curve entirely and calculates frictional power directly from the measured friction torque. This approach provides a more direct and potentially more accurate estimation of the heat source, but it is only applicable to the tested sealing system. In contrast, the Gümbel curve can be used to estimate frictional power of similar sealing systems. Estimation und validation of the contact temperature for RSS optimizations As previously discussed, several optimization strategies have been implemented to enhance the performance of the radial shaft sealing system, each targeting specific components of the system. The modifications summarized in Table 1 are primarily aimed at reducing the contact temperature through systematic adjustments to the RSS. The RSS modification number reflects the sequential application of these measures, beginning with a baseline configuration consisting of a standard seal design (BAUM), a standard garter spring (STD), and the elastomer 75 FKM 585. Subsequent modifications include: • Modified spring to reduce radial load (MOD) • Selection of a high-performance elastomer (75 FKM 170055) • Friction-optimized sinusoidal sealing profile (PSS) These design changes are intended to mitigate thermomechanical stress at the sealing gap and to significantly improve system reliability and efficiency, particularly under high circumferential speed conditions. The estimation of the expected contact temperature, as well as the subsequent validation experiments, was carried out using similar test parameters as those used for measuring the friction torque and determining the Gümbel curve. The applied test parameters are listed in Table 2. Various circumferential speeds were investigated. Prior to each test, the oil sump was preheated to 70 °C and subsequently not actively cooled, to allow self-heating due to the friction. Since direct temperature measurement at the sealing contact is not possible, the measured oil sump temperature can be compared with the theoretically estimated oil sump temperature derived from the approaches of the ExACT Equation and Thermal Net- Science and Research 67 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0022 Condition Value Unit Speed 1000, 1500, 2000, 2500, 3000 rpm Circumferential Speed 6.3, 9.4, 12.6, 15.7, 18.8 m/ s Medium Mineral gear oil ISO VG 220 Temperature 70 - self heating °C Cycle 20/ 4 - 20h speed / 4h pause Duration 504 h n 2 RSS Table 2: Overview of the test conditions for temperature estimation and validation RSS-Modification-No. RSS Elastomer Spring Dimension 1 BAUM 1 75 FKM 585 STD 3 120-150-12 2 BAUM 1 75 FKM 585 MOD 4 120-150-12 3 BAUM 1 75 FKM 170055 MOD 4 120-150-12 4 PSS 2 75 FKM 585 - 120-150-12 1 BAUM = Standard Seal Design 2 PSS = Premium Sine Seal 3 STD = Standard garter spring 4 MOD = Modified garter spring for reduced radial load Table 1: Summary of Radial Shaft Seal (RSS) optimization strategies Complementary to the oil sump temperature analysis, Figure 9 shows the estimated contact temperature for the same range of RSS modifications and speeds. The results indicate a consistent increase in contact temperature with rising speed as well. The estimations indicate that the baseline configuration at a speed of 18.8 m/ s leads to a contact temperature of approximately 200 °C. This value significantly exceeds the permissible temperature limits for both the lubricant and the elastomer, representing a critical thermal load for both components. Furthermore, it can be observed that the applied optimizations measures lead to a significant reduction in contact temperature up to 20 °C consistently across all estimation methods. Science and Research 68 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0022 work. This enables improved validation of the contact temperature estimation. Figure 8 shows the estimated oil sump temperature as a function of the RSS modification across the investigated speeds. The figure includes three estimation methods, complemented by measured temperatures obtained from the friction torque tests. The results indicate that the experimentally determined temperatures lie between the calculated values, suggesting a realistic approximation by the models. Furthermore, a clear increase in temperature with higher speed is observed, along with a temperature reduction of up to 20 °C due to the applied modification measures. Figure 8: Estimated oil sump temperature results for various modifications under different circumferential speeds - friction torque test Figure 9: Estimated contact temperature results for various modifications under different circumferential speeds - friction torque test This correlation is also reflected in the temperature estimations derived from the long-term validation tests, see Figure 10. In this case, the estimation was only based on the Gümbel curve. For the baseline configuration, a contact temperature of approximately 200 °C is estimated, confirming the previously identified critical thermal load. However, through the implemented RSS modification measures, the temperature can be significantly reduced - depending on the method, by approximately 30 to 70 °C. A similar trend is observed in the estimation of the oil sump temperature. Notable, the predicted values align well with the measurements, supporting the validity of the contact temperature estimations. The contact temperatures estimated during the test series are clearly confirmed by the analysis of the sealing edges. Figure 11 shows the sealing edges of the different optimized RSS configurations after testing at 18.8 m/ s. In the baseline configuration, significant hard deposits are visible at the sealing edge, indicating thermal degradation of the lubricant. Additionally, the elastomer shows discoloration toward both the air-side and oilside contact area, further suggesting thermal overload. Leakage was observed as well. The radial load optimized RSS variant, with a modified spring, shows significant reduction in deposits and a narrower seal wear band width. No leakage occurred in this test. This optimization supports the previously estimated contact temperature, as the reduced spring load results in a temperature decrease of approximately 30 °C to 125 °C. Another measure, the use of the high-performance elastomer 75 FKM 170055, lead to another substantial reduction in deposits and seal wear. The switch to the PSS profile demonstrated a significant improvement in terms Science and Research 69 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0022 Figure 10: Comparison between estimated contact and oil sump temperatures and the measured oil sump temperature over the 504-hour long-term validation test Figure 11: Seal wear band resulting from the applied modifications after 504 hours validation tests at 18.8 m/ s Post-test analysis of the sealing edges supports these findings: while the baseline configuration shows significant deposits, discoloration and leakage, the optimized variants clearly show reduced deposit formation, narrow wear band width, and no leakage. In particular, the highperformance elastomer 75 FKM 170055 and the PSS profile demonstrated the lowest estimated contact temperatures for the long-term validation test and the best overall condition after test. A promising outlook arises from the potential to estimate the model-based contact temperature estimation based on some basic measurements. It enables early assessment of sealing system functionality and predictive analysis of thermal and tribological stresses - before critical conditions such as leakage or material failure occur. As a result, the diagnostic value of test data is significantly enhanced, supporting the development of robust and high-performance sealing systems for high-speed applications. Reference [1] Freudenberg FST GmbH [2] Feldmeth, S.: Simulative Bestimmung der Temperatur im Dichtkontakt von Radial-Wellendichtungen. Dissertation, Universität Stuttgart, 2025. [3] Bohnert, C.; Heilemann, J.; Thielen, S.: Th ermische Simulation Dichtkontakt. FVA-Projekt 574 IV, Satus: Vorläufiger Abschlussbericht, Forschungsvereinigung Antriebstechnik e.V. (FVA), 2025. Science and Research 70 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0022 of deposits formation and seal wear. This profile also exhibited the lowest estimated contact temperature for the validation test. Summary and Outlook Radial shaft sealing systems used to seal rotating components face significant thermal and tribological challenges at high circumferential speeds. These stresses can lead to degradation of the lubricant and damage of the elastomer, resulting in premature failure of the sealing system. The aim of this paper was to reduce the thermomechanical load and improve the performance of the system through targeted modifications of the RSS, including spring modification, selection of high-performance elastomer, and optimization of the sealing-lip profile. To evaluate the effectiveness of these optimization measures, model-based approaches were used to estimate contact and oil sump temperatures - specifically the ExACT [2] equation and a Thermal Network [3]. Both methods enable reliable predictions of temperature development in the sealing contact and were validated using experimental data. The results show that the baseline configuration can reach a critical temperature of up to 200 °C at high speed (18.8 m/ s), exceeding the thermal limits of both lubricant and elastomer. Through the applied optimization measures, the contact temperature was reduced by up to 70 °C depending on the modification - an effect confirmed both by contact temperature model and experimental validation. 1 Introduction and problem statement Approximately 70 % of all failures and malfunctions in hydraulically operated machines can be traced back to contamination of the hydraulic fluid where solid particles represent the most frequent cause of malfunctions. They lead to material abrasion and increased wear on pumps, motors, valves, and cylinders, thereby significantly reducing the durability of these components. A distinction can be made between internally generated and externally introduced particles. Internally generated particles arise e.g., from friction, corrosion, or cavitation within the hydraulic system. Typical examples include wear particles made of steel from pumps and motors, brass or bronze from sliding bearings, and elastomers from seals or guiding elements. Externally introduced particles enter the system from the environment, for instance sand passing through defective seals or entering during maintenance activities. Particularly small and hard particles have the potential to cause severe damage, since the clearance in hydraulic components is usually less than 10 µm. The concentration of such small particles in hydraulic oil is especially high [1] causing problems. The consequences of wear processes are particularly severe in construction machinery, agricultural and forestry machinery, industrial trucks, and machine tools. Undetected wear can lead to unplanned downtime, production losses, and considerable costs. With the increasing use of proportional and servo hydraulics, the sensitivity of hydraulic systems to sustain damage is increasing [2, 3]. Currently, the condition of hydraulic fluids is monitored by taking oil samples from the machines and analysing them in a laboratory. While these analyses provide detailed results, they are time-consuming, costly, and only reflect the condition at the time of sampling. Inline particle counters represent an alternative, as they continuously measure the number and size of particles in the oil. However, these devices often detect water and air bubbles wrongly as solid particles distorting the measurement results. Furthermore, these devices do not capture particle shape, although this parameter provides crucial information about material, wear mechanisms, and possible sources of origin. In recent years, AI based systems have been developed to automatically detect and classify particles in oil [4, 5]. In [4] a system for detecting and classifying wear particles in engine oils was presented that assigns them to five categories: cutting wear particles, sliding wear particles, fatigue wear particles, spherical particles and nonmetallic particles. The training data required for this system were obtained directly from real oil samples by using an optical particle sensor that extracted oil from Science and Research 71 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 Concept for the generation of single-type wear particles for training the AI-based image processing of a particle sensor Dennis Jess, Andreas Ligocki* Presented at GfT Conference 2025 Most failures of hydraulically operated machines are caused by solid particles that remove material from components within the hydraulic circuit. If wear remains undetected, unplanned downtime and high costs can result. The HydroVision project therefore develops an AI-based particle sensor that detects, counts, and classifies particles in hydraulic oil, enabling interpretation of wear processes. Suitable training data is required, yet previous approaches are unsuitable because they do not allow generation of single-type training data. Consequently, the relationship between particle shape and material cannot be clearly established, and data labelling is hindered. This paper presents a concept for generating singletype training data. In addition, results from oil particle analyses are presented, which form the basis for defining the fine specifications of the particle sensor. Keywords Hydraulic oil; Wear particles; Particle sensor; Image processing; Artificial intelligence; Training data generation; Predictive maintenance Abstract * M. Eng. Dennis Jess (corresponding author) Prof. Dr.-Ing. Andreas Ligocki Ostfalia University of Applied Sciences Salzdahlumer Str. 46/ 48 38302 Wolfenbüttel, Germany Regional Development Fund (ERDF), Ostfalia University of Applied Sciences in Wolfenbüttel in cooperation with a medium-sized company is pursuing an alternative approach. The objective is likewise the development of a particle sensor that, using optical measurement technology and AI-based image processing, detects the smallest particles in oil circuits, counts them, determines their contours and shapes, and classifies them into different particle types. For training the AI-based image processing, exclusively single-type particle samples are generated to reliably investigate the relationship between particle shape and material and to significantly simplify data labelling. This approach allows better interpretation of wear phenomena in mechanical components of hydraulic systems in construction machinery, agricultural and forestry machinery, industrial trucks, and machine tools, and planning maintenance and repair measures more effectively. In the following, the project-specific concept for generating single-type particles is presented. After this the results from the analysis of oil particles, which form the basis for defining the fine specifications of the particle sensor as well as the requirements for the training data are discussed. 2 Phases of the HydroVision project The HydroVision project is divided into five phases, hereinafter referred to as STEPs. These comprise particle analysis and the establishment of a particle database, the development of test benches, the generation of training Science and Research 72 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 the main lubrication circuit of an engine and captured images of the particles contained in it. From these images individual particle images were extracted and manually labelled. In [5] an AI based system for monitoring marine lubricants was developed that records size, shape, and number of wear particles to assess the condition of hydraulic systems. For this purpose, oil samples were artificially prepared by adding defined quantities of iron, copper, and aluminium particles to the oil. The samples were then passed through a transparent detection chip and recorded under a microscope using a high-resolution camera. Again, the individual particle images were extracted from the recordings and manually annotated. However, both approaches reveal significant weaknesses: In [4], no separation of particles in the oil samples was carried out to create single-type particles, which restrained the establishment of a relationship between particle shape and material. This made data labelling considerably more difficult and reduced the quality of the datasets. In [5], single-type separation was also lacking. Although in principle the relationship between shape and material could have been investigated, the materials were introduced into a single oil sample instead of being examined in separate samples. In addition, only metallic particles were considered, while non-metallic particles such as elastomers, sand particles, or dust were not considered. As part of the three-year HydroVision research project funded by the N-Bank with resources from the European Figure 1: Overview of the phases of the HydroVision project data and development of AI, the realisation and testing of a prototype, and finally the optimisation of the prototype. The schematic workflow of the HydroVision project is shown in Figure 1. A particular focus of this chapter is the presentation of the concept for generating single-type training data, which is primarily implemented in the first three project phases and forms the basis for the development of the particle sensor. STEP 1 - Particle analysis and establishment of a particle database In the first project phase, fundamental investigations are carried out to assess the conditions of hydraulic systems from different machines in real application cases and target customer segments. For this purpose, oil samples are taken, and the contained particles are analysed under a light microscope. The particles are counted and characterised according to their size, shape, and material. All collected data are documented in a central particle database. Based on this database, the particle types to be detected by the sensor under development are defined. The database thus forms the basis for the artificial and single-type generation of particles. Finally, the fine specifications of the particle sensor are defined based on the identified particle types. STEP 2 - Development of test benches In this phase, the focus lies on the development of various test benches. They serve both to generate training data for the AI and to provide an environment for its training and testing. In total, three test benches are set up. 1.) Particle generation test bench: A compact, modular test bench is designed, engineered, and built to artificially generate single-type wear particles from defined materials with specified sizes and shapes, corresponding to the previously defined particle types. This allows the production of e.g., metallic particles resulting from gear or bearing wear, as well as elastomer particles typically generated by the wear of seals or guiding elements in hydraulic cylinders. 2.) Dry test bench: Afterwards, a dry test bench is designed, engineered, and built. It consists of a linear axis and the components of the Opti module. This module includes a CCD camera and a light source arranged opposite each other. The generated particles are fixed on a movable slide, which is guided along the linear axis between the camera and the light source, thereby simulating the flow in a hydraulic system. During this process, images of the particles are captured. The dry test bench eliminates the influence of hydraulic oil and particle movement. Furthermore, in the initial operating phase, it allows rapid adjustment and easy replacement of all components. Figure 2 shows the schematic setup of the dry test bench. 3.) Wet test bench: Finally, a wet test bench is designed, engineered, and built. It consists of several isolated miniature hydraulic circuits in which the oil is circulated by an external peristaltic pump. These circuits are deliberately loaded with the generated particles. The Opti module (camera and light source) can be flexibly connected to the circuits, enabling images of particles to be captured in the oil flow under realistic conditions. This setup ensures that single-type particles can be examined without cross-contamination. Figure 3 shows the schematic setup of the wet test bench. Science and Research 73 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 Figure 2: Schematic setup of the dry test bench Figure 3: Schematic setup of the wet bench A compact housing with an integrated measurement cell for particle flow is constructed and manufactured for the sensor components. In addition, a service concept for replacing selected components is developed. A study examines suitable installation positions within the machine’s hydraulic circuit to ensure that particle detection is not impaired. Finally, the prototype is tested in a field trial under real operating conditions, and its performance is evaluated. STEP 5 - Optimisation of the prototype In the final phase, the functionality of the particle sensor is validated and optimised. The evaluation during the field test focuses on the extent to which the specified particles can be detected, counted, and assigned to their potential source. Based on these results, the operational limits of the sensor are fixed, and application scenarios are derived. Furthermore, the collected particles provide the foundation for developing precise recommendations for maintenance. 3 Particle analysis Particle analysis represents the first phase of the Hydro- Vision project and is divided into four steps including oil sampling, membrane filter preparation, particle data acquisition, and documentation and evaluation. The objective of the particle analysis is to use the collected data Science and Research 74 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 STEP 3 - Generation of training data and development of AI The dry and wet test benches are equipped with the optical components and put into operation in this step. The single-type particles generated in the particle generation test bench are introduced into both test benches. In an iterative process, the test benches, optical setup, image acquisition software, and AI-based image processing are continuously refined. After the hardware and software have been verified, particle images are acquired, labelled, and compiled into training and test datasets. Based on these datasets, suitable AI models are trained, tested, and optimised. The insights gained are then used to further adapt the optical setup and the image acquisition software. Figure 4 illustrates different variations of the optical setup with alternative arrangements of light source and CCD camera, variable light sources in terms of wavelength and intensity, as well as the use of lens systems. This approach enables the continuous improvement of the test benches, optical setup, image acquisition software, and AI-based image processing. STEP 4 - Implementation and testing of a prototype The first functional prototype is developed based on the previous results. The prototype is designed as a plugand-play solution that can be integrated into standard hydraulic circuits. Figure 4: Variations of the optical measurement to define the relevant particle types and the fine specifications for the particle sensor to be developed. 3.1 Oil sampling As part of the sampling process, oil samples were taken from forklift trucks, agricultural machines, forestry machines, construction machines, and machine tools. The sampling was carried out in accordance with the guidelines of Oelcheck GmbH [6] and Bureau Veritas GmbH [7] to ensure the representativeness of the oil samples and to enable comparative analysis assignments. The oil samples were extracted from the tank of each machine using a sampling pump. The extraction was performed from the mid-level of the oil, to avoid disturbances from air at the surface or particles deposited at the bottom. At least five oil samples with a volume of 100 ml were taken from each machine. One sample was analysed in the in-house laboratory, another in an external laboratory, and the remaining samples were kept in stock for potential repetition analyses. Each oil sample was assigned a unique identification number. A sampling protocol documented not only the machine data (e.g. machine type, operating hours) but also oil data (e.g. manufacturer, designation) and sampling data (e.g. sampling date, sample volume). In total, oil samples were taken from three CNC lathes, one CNC milling machine, one excavator, two wheel loaders, three forklift trucks, one portal lifting truck, one forest crawler vehicle, two harvesters, and one tractor. Table 1 provides an overview of the machines from which oil samples were taken. 3.2 Preparation of membrane filters The preparation of the membrane filters for the analyses was carried out according to ISO 4407 [8] using a vacuum filtration system. Figure 5 shows the setup of the ap- Science and Research 75 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 Figure 5: Setup of the vacuum filtration apparatus p No. Machine type Machine manufacturer Machine name Operating hours [h] Oil usage time [h] 1 CNC lathe Taiwan TAKISAWA NEX-108M 41 066 unknown 2 CNC lathe Taiwan TAKISAWA FX-800 27 243 unknown 3 CNC lathe DMG MORI CLX450 3 252 3 252 4 CNC milling machine MAKINO a51nx 41 264 unknown 5 Excavator Öswag Eurocat 140 HVS 3 248 264 6 Wheel loader Schäffer 2026 1 448 143 7 Wheel loader Liebherr L 506 Compact 807 807 8 Forklift truck Hyster-Yale J2.0XN MWB 12 283 11 878 9 Forklift truck Hyster-Yale H3.5FT 3 796 440 10 Forklift truck Hyster-Yale H3.5FT 3 906 1 039 11 Portal lifting truck Combilift MG3796-05 8 339 1 918 12 Forest crawler vehicle Agria-Werke AGRIA 9700e 134 134 13 Harvester Ponsse Ergo A090247 10 681 10 681 14 Harvester Ponsse Ergo A090853 9 828 925 15 Tractor Deutz Agrotron 150 TT3 6 692 700 Table 1: Overview of machines from which oil samples were taken After microscopic analysis, particle characteristics (e.g., area, maximum diameter, minimum diameter, brightness) were exported as CSV files, and the recorded particle images as PNG files from the microscope software. The determined particle counts were recalculated according to DIN 51455 [9] to the total filter area and, if necessary, to a sample volume of 100 ml. Microscope settings and particle size distributions according to ISO 16232 [10] and ISO 4406 [11] were documented in the accompanying protocol. 3.4 Documentation and evaluation All information obtained in the previous phases was documented in a central particle database. The database contains details on the sample, the associated machine and the oil used, information on the selected analysis parameters and results, as well as the characteristics of the individual particles. Table 2 provides an overview of the stored information and the respective data fields. All entries are linked via the sample ID, ensuring complete traceability from general sample information down to the characteristics of individual particles. Particle size distribution The determination of the particle size distribution in the machines was carried out according to the methodology described in Section 3.3. Table 3 provides an overview of the total number of particles detected and their percentage distribution across the individual size classes. The particle size was defined as the respective maximum diameter. Table 3 shows that most particles are in the size range of 5.1 to 15 µm, with an average percentage of 68.6 %. As Science and Research 76 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 paratus used. The filter medium was a polyester membrane filter with a diameter of 47 mm and a pore size of 1 µm. A separate membrane filter was inserted into the vacuum filtration system for each oil sample. To ensure a uniform particle distribution, the oil sample was first homogenised. The exact sample volume was then measured and documented before being poured into the funnel of the filtration unit. The negative pressure generated in the suction bottle drew the oil sample through the membrane filter, causing the particles to be deposited on its surface. Filtration continued until the membrane filter was completely dry. The prepared membrane filter was labelled with the identification number of the oil sample and stored in a clean Petri dish until microscopic analysis. 3.3 Acquisition of particle data The microscopic particle analysis was carried out in accordance with DIN 51455 [9] and ISO 16232 [10]. The particles deposited on the membrane filters were analysed using the automated VHX-7000 digital microscope (Keyence Deutschland GmbH). The setup of the VHX-7000 digital microscope is shown in Figure 6. According to DIN 51455 [9], the analysis was performed in reflected light mode within a circular area of 35 mm in diameter at the centre of the membrane filter, with a pixel resolution of < 1 µm per pixel. The acquisition of particle data followed ISO 16232 [10], as this standard considers both light and dark particles. Illumination adjustment, particle counting, particle size classification, acquisition of particle characteristics, and particle image capture were performed automatically. Figure 6: Setup of the VHX-7000 digital microscope particle size increases, the number of particles decreases sharply. Only about 0.9 % of all particles are larger than 200 µm. Since the proportion of particles above 200 µm is negligible and mainly consists of fibres, the upper size limit for particle counting is set at 200 µm. As the particle sensor to be developed must determine the oil contamination level according to ISO 4406 [11], like conventional particle counters, the lower size limit for particle counting is set at 4 µm. It is also noticeable that machines with a high number of particles, such as the CNC milling machine (No. 4), the excavator (No. 5), and the tractor (No. 15), show a higher percentage of particles in the size ranges 15.1 to 50 µm. This confirms the assumption in [12] that with increasing operating time and progressive wear, a greater number of larger particles is generated. On this basis, the lower size limit for particle shape recognition is set at 20 µm, since the progression of wear can be reliably assessed from larger particles. Particle materials Figure 7 illustrates example images of typical metal, sand, elastomer, and fibre particles obtained from the analysed machines. In addition, their characteristic optical and morphological features are listed in bullet points. Metal particles are often shiny and exhibit sharp contours. Sand particles appear matte, with sharp edges and a colour range from light to dark brown. Elastomer particles are matte, with soft contours, and often display irregular surfaces or pores. Fibres, in contrast, usually appear elongated and threadlike. Table 4 shows the particle materials detected in the analysed machines. These include steel (St), zinc (Zn), alu- Science and Research 77 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 Table sheet Contained data fields Sample information table • Sample data: Sample ID, sampling time, sample volume (ml), sampling location • Machine data: Machine type, machine manufacturer, machine designation, operating hours (h), oil usage time (h), application field • Oil data: Oil manufacturer, oil designation, oil type, kinematic oil viscosity (cSt) • Environment: Environmental influences Sample results table • Microscope data: Microscope manufacturer, microscope designation • Filter data: Diameter (mm), extraction area (mm²), flow-through area (mm²), pore size (μm) • Image acquisition parameters: Illumination type, pixel resolution (μm/ pixel), grayscale threshold values • Particle size distribution: Particle size distribution according to ISO 16232 and ISO 4406 Particle data table • Particle information: Particle ID, particle image name • Particle geometry: Maximum diameter (μm), minimum diameter (μm), aspect ratio, area (μm²) • Optical properties: Average brightness (%) Table 2: Overview of tables in the particle database No. Machine type Machine name Total particle count Particle size distribution [%] 5.1 - 15 [μm] 15.1 - 25 [μm] 25.1 - 50 [μm] 50.1 - 100 [μm] 100.1 - 200 [μm] > 200 [μm] 1 CNC lathe NEX-108M 1 226 69.2 16.9 9.2 3.0 0.5 1.1 2 CNC lathe FX-800 3 017 75.8 16.8 5.7 1.0 0.2 0.5 3 CNC lathe CLX450 1 002 71.5 14.6 8.6 3.6 0.5 1.3 4 CNC milling machine a51nx 7 631 53.3 20.8 19.5 5.2 0.5 0.7 5 Excavator Eurocat 140 HVS 6 447 57.3 23.9 16.3 1.9 0.2 0.4 6 Wheel loader 2026 3 067 82.4 10.3 6.0 0.7 0.1 0.5 7 Wheel loader L 506 Compact 1 379 72.6 14.4 8.6 2.4 0.6 1.5 8 Forklift truck J2.0XN MWB 1 723 63.4 17.2 13.3 4.1 0.5 1.3 9 Forklift truck H3.5FT 1 323 67.6 15.0 12.5 3.4 0.4 1.1 10 Forklift truck H3.5FT 1 699 70.6 14.4 10.3 3.1 0.9 0.8 11 Portal lifting truck MG3796-05 1 114 77.9 17.8 3.8 0.4 0.0 0.9 12 Forest crawler vehicle AGRIA 9700e 3 858 93.3 4.3 1.6 0.3 0.1 0.4 13 Harvester Ergo A090247 1 693 63.6 17.0 12.6 4.8 1.0 0.9 14 Harvester Ergo A090853 1 338 58.4 17.1 16.7 4.8 1.3 1.6 15 Tractor Agrotron 150 TT3 32 836 52.0 21.0 19.6 6.0 0.8 0.6 Average of all machines 68.6 16.1 11.0 3.0 0.5 0.9 Table 3: Overview of machines and particle size distributions the tractor (No. 15). Fabric fibres were present in all oil samples. These results reinforce that the particle sensor to be developed must be capable of distinguishing between metal, sand, elastomer, and fibre particles to reliably interpret wear processes. However, the composition of the particle material spectrum strongly depends on the design and operating conditions of the respective hydraulic system. Science and Research 78 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 minium (Al), brass (Ms), bronze (Br), sand (Sd), elastomers (El), and fabric fibres (Fb). Steel and aluminium particles were detected in all machines as illustrated in Table 4. Zinc, brass, and bronze particles appeared only sporadically. Sand particles were found in about half of the machines, with a significantly higher proportion observed in the excavator (No. 5) and Particle image Magnification × 800 × 800 × 800 × 1 000 Maximum diameter 72 μm 91 μm 84 μm 70 μm Material Metal (Steel) Metal (Zinc) Metal (Aluminium) Metal (Brass) Features • black, reddish • matte, slightly shiny • sharp edges • light gray, silvery, bluish • shiny • sharp edges • light gray, silvery • shiny • sharp edges • golden • shiny • sharp edges Particle image Magnification × 800 × 800 × 800 × 300 Maximum diameter 68 μm 61 μm 68 μm 234 μm Material Metal (Bronze) Sand Elastomer Fabric Fibre Features • brownish • shiny • sharp edges • light to dark brown • matte • sharp edges • matte • soft edges • often contains holes • light to dark • matte • elongated and thin Figure 7: Example images of particles from different materials p No. Machine type Machine name St Zn Al Ms Br Sd El Fb 1 CNC lathe NEX-108M ✔ - ✔ - - - ✔ ✔ 2 CNC lathe FX-800 ✔ - ✔ - - - ✔ ✔ 3 CNC lathe CLX450 ✔ - ✔ - - - - ✔ 4 CNC milling machine a51nx ✔ - ✔ - - - ✔ ✔ 5 Excavator Eurocat 140 HVS ✔ ✔ ✔ - - ✔ ✔ ✔ 6 Wheel loader 2026 ✔ - ✔ - - ✔ - ✔ 7 Wheel loader L 506 Compact ✔ - ✔ - - ✔ - ✔ 8 Forklift truck J2.0XN MWB ✔ - ✔ - - ✔ ✔ ✔ 9 Forklift truck H3.5FT ✔ - ✔ - - - ✔ ✔ 10 Forklift truck H3.5FT ✔ ✔ ✔ - ✔ ✔ - ✔ 11 Portal lifting truck MG3796-05 ✔ - ✔ - - - ✔ ✔ 12 Forest crawler vehicle AGRIA 9700e ✔ - ✔ - - - - ✔ 13 Harvester Ergo A090247 ✔ - ✔ - ✔ ✔ ✔ ✔ 14 Harvester Ergo A090853 ✔ - ✔ - - ✔ ✔ ✔ 15 Tractor Agrotron 150 TT3 ✔ ✔ ✔ ✔ ✔ ✔ ✔ ✔ Table 4: Overview of the machines and the particle materials identified Shape distribution of metallic particles According to [13], metallic particles can be classified into the basic shapes: lamella, splinter, chip, and sphere. Table 5 summarises the characteristic features and formation mechanisms of these shapes. Figure 8 shows exemplary images of lamellar, splinter, and chip-shaped particles illustrating the differences between these particle shapes. According to the earlier defined restrictions, only particles with a size of 20 µm or larger were considered for shape classification. Not every machine provided enough analysable particle images. Therefore the evaluation was limited to the five machines with the highest number of particles in the size range of 20 to 200 µm. From each of these machines, 100 images of metallic particles were randomly selected, analysed, compared with the features listed in Table 5, and their particle shape was determined. Table 6 provides an overview of the Science and Research 79 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 Particle shape Features Formation Lamella • flat surface • scratch marks on the surface • large surface area relative to thickness • small thickness • irregular edges • sharp edges (compared to splinters less sharp) • surface fatigue (e.g. pittings) • detachment of material layers • repeated loading Splinter • rough surface • irregular edges • many sharp edges • fracture due to overload • crack formation and spalling • impact or shock loading Chip • elongated shape • fibrous, ribbonor needle-shaped • curved or spiral • sharp edges • may have pointed ends • cutting or abrasive wear • sharp edges on hard surfaces penetrating into soft surfaces Sphere • round or approximately spherical • local melting and solidification (e.g. welding, grinding) Table 5: Overview of particle shapes with their features and formation mechanisms Particle image Magnification × 800 × 800 × 1 000 Maximum diameter 68 μm 68 μm 75 μm Material Zinc Aluminium Aluminium Shape Lamella Splinter Chip Features • flat shape • large surface area relative to thickness • less sharp edges compared to splinters • rough surface • irregular shape • many sharp edges and fracture lines • elongated, fibrous or ribbon-like shape • partly curved • thin • sharp edges Figure 8: Example images of particles of different shapes No. Machine type Machine name Particle shape distribution [%] Lamella Splinter Chip Sphere 1 CNC milling machine a51nx 39 37 24 0 2 Excavator Eurocat 140 HVS 58 18 24 0 3 Forklift truck J2.0XN MWB 69 6 25 0 4 Harvester Ergo A090247 58 21 21 0 5 Tractor Agrotron 150 TT3 56 28 16 0 Average of all machines 56 22 22 0 Table 6: Overview of the machines and the shape distributions of metallic particles lamellar, splinter, and chip-shaped particles to reliably interpret wear processes of metallic components providing a basis for developing a particle sensor that enables reliable condition monitoring of hydraulic systems. In the next step, the requirements for the particle generation test bench will be defined. It will then be designed and built. Subsequently, initial experiments will be carried out to produce particles from selected materials with defined sizes and shapes. By varying the experimental parameters, size and shape of the particles can be specifically influenced. The objective is to identify parameter combinations that allow the production of realistic particles. The generated particles will be examined under a light microscope and compared with particles from real oil samples to verify their similarity. These artificially created, realistic single-type particles will be applied to the dry and wet test benches, forming the basis for the image generation (training data for the sensor under development) using the Opti module. References [1] Karberg & Hennemann GmbH & Co. KG: Ratgeber Öl, 2023. Online verfügbar unter: https: / / www.cjc.de/ wp-content/ uploads/ 2023/ 09/ ratgeber -oel_de.pdf (accessed on August 18, 2025) [2] M. Jocanović, S. Andrić, M. Lazarević, D. Lukić: Example of Good Maintenance Practice for Maintaining the Health of a Hydraulic System. In: M. Rackov, R. Mitrović, M. Čavić (eds), Machine and Industrial Design in Mechanical Engineering. KOD 2021, Mechanisms and Machine Science, vol. 109, Springer, Cham, 2022. https: / / doi.org/ 10.1007/ 978-3-030-88465-9_36 [3] V. Karanovic, M. Jocanović, S. Lalos, B. Z. Knezevic: Oil Cleanliness Class Influence on Wear Intensity of Piston- Cylinder Contact Pair Inside of Hydraulic Distribution Valve, DEMI 2015, 12th International Conference on Accomplishments in Mechanical and Industrial Engineering, Banja Luka, May 2015. [4] S. Fan, T. Zhang, X. Guo, A. Wulamu: FFWR-Net: A feature fusion wear particle recognition network for wear particle classification. Journal of Mechanical Science and Technology, vol. 35, no. 4, pp. 1699-1710, Springer, 2021. https: / / doi.org/ 10.1007/ s12206-021-0333-6 [5] C. Bai, J. Ding, H. Zhang, Z. Xu, H. Liu, W. Li, G. Li, Y. Wei, J. Wang: Research on Abrasive Particle Target Detection and Feature Extraction for Marine Lubricating Oil. Journal of Marine Science and Engineering, 12, 677, MDPI, 2024. https: / / doi.org/ 10.3390/ jmse12040677 [6] OELCHECK GmbH: Ölprobenentnahme. Online verfügbar unter: https: / / de.oelcheck.com/ fileadmin/ user_ upload/ gebrauchsanleitungen/ Anleitung_Allgemein.pdf (accessed on August 18, 2025) [7] Bureau Veritas GmbH: Probenentnahme mit der Probenpumpe, 2020. Online verfügbar unter: https: / / oil-testing.de/ wp-content/ uploads/ 2020/ 10/ BV- Probenentnahme-mit-der-Probenpumpe.pdf (accessed on August 18, 2025) Science and Research 80 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 examined machines and the shape distribution of metallic particles. Table 6 shows the erratic shape distribution of the metallic particles. The largest proportion consists of lamellar particles, followed by splinter and chip particles, which occur at approximately the same frequency. Spherical particles could not be detected. This demonstrates that the particle sensor under development must be capable of reliably identifying and distinguishing lamellar, splinter, and chip particles to accurately interpret wear processes. Furthermore, the results highlight that real oil samples are only of limited suitability for training AI models, as the shape distribution of metallic particles is highly unbalanced. To ensure that the AI can reliably recognize all particle shapes, a balanced representation of particle shapes in the training dataset is required. 4. Conclusion and outlook As the objective of the HydroVision research project, a particle sensor is being developed that uses optical measurement technology and AI-based image processing to detect, count, and classify particles in hydraulic oils, while also identifying their shapes and contours. For the realisation of the AI-based image processing, suitable training data must first be generated. Previous approaches [4, 5], however, did not allow the generation of single-type training data, preventing a clear link between particle shape and material, and thus making data labelling considerably more difficult. This paper therefore presented a concept that enables the production of single-type training data. This approach allows the targeted investigation of the relationship between particle shape and material, it simplifies data labelling, and ensures a high data quality. Particles from real oil samples of construction machinery, agricultural and forestry machinery, industrial trucks, and machine tools were analysed within the project. Based on these analyses, the relevant particle types were exposed, the functional fine specifications of the particle sensor under development were defined and the foundation for implementing the proposed concept for generating single-type training data was laid. From the analysis results, this indicates that real oil samples are only of limited suitability for training AI models, since the shape distribution of metallic particles is highly unbalanced. To ensure that the AI can reliably recognize all particle shapes, a balanced distribution within the training dataset is required. The sensor under development must be able to detect and count particles in the size range of 4 to 200 µm. Particles larger than 20 µm must be identified by shape and classified into the categories metal, sand, elastomer, and fibre. Metal particles must additionally be classified into [8] ISO 4407: Hydraulic fluid power - Fluid contamination - Determination of particulate contamination by the counting method using an optical microscope. Geneva: International Organization for Standardization, 2002. [9] DIN 51455: Flüssige Mineralölerzeugnisse - Bestimmung der Partikelanzahl und Partikelgröße in Ölen. Berlin: Deutsches Institut für Normung, 2020. [10] ISO 16232: Road vehicles - Cleanliness of components of fluid circuits. Geneva: International Organization for Standardization, 2018. [11] ISO 4406: Hydraulic fluid power - Fluids - Method for coding the level of contamination by solid particles. Geneva: International Organization for Standardization, 2021. [12] B. Bhushan (ed.): Modern Tribology Handbook, 2 Vols. CRC Press, Boca Raton, 2000. https: / / doi.org/ 10.1201/ 9780849377877 [13] F. Bauer (ed.): Tribologie - kompakt und praxisnah. 1 st edition, Springer Fachmedien Wiesbaden (Springer Vieweg), Wiesbaden, 2021. https: / / doi.org/ 10.1007/ 978-3-658- 32920-4 Science and Research 81 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 DOI 10.24053/ TuS-2025-0023 News 82 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 The Austrian Tribology Society (ÖTG) is strongly committed to promoting all areas of tribology - from research and development to practical industrial applications. This also includes the active support of new companies, particularly during their founding phase and early development. In this context, the ÖTG serves as an initiator and founding platform. Today, the scientific association holds shares in three companies that together offer a wide range of services in tribology, material development and characterization, as well as modeling, simulation, and digitalization - all based at the Technology and Research Center in Wiener Neustadt, around 40 kilometers south of Vienna. Österreichische Tribologische Gesellschaft A scientific association as an incubator for new businesses AC2T research GmbH (AC2T) is the Austrian Centre of Competence for Tribology, established in 2002. Since then, AC2T has evolved into one of the largest privately operated R&D service providers in the field of tribology worldwide. AC2T has extensive and unique infrastructure for carrying out highly interdisciplinary projects. Thereby, the Centre builds on its core competencies: Tribological system, material, surface, and lubricant characterization; chemical analysis; measuring technology and sensor development; modelling, simulation and data science. A team of 100 people develops customized solutions in the fields of transportation, energy, production, and medicine:  Friction-optimized systems: Design and optimization of surfaces and materials to reduce friction and increase the reliability and energy efficiency of mechanical components.  Wear reduction strategies: Determination of the service life behavior of tribological components to increase robustness and replace critical raw materials.  Sustainable lubrication: Development of lubricants and lubrication strategies up to online monitoring systems, taking sustainability and circular economy into account.  Synaptic tribology: Combining tribological expertise with data science, simulation and digitalization to make tribological systems and their complex behavior in practice predictable. Phone: +43 2622 81600-0 | E-mail: office@ac2t.at | Web: https: / / www.ac2t.at/ The ÖTG is pleased to present its three success stories: News 83 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 Aerospace & Advanced Composites GmbH (AAC) is a research and technology company in the field of materials testing and development for special applications in aerospace, aviation, energy, industry and defense, founded in 2010. AAC supports international customers and partners - including ESA - throughout the entire value chain in the development, testing, and qualification of durable, efficient, and safe solutions:  Material characterization and tribological testing under extreme conditions, such as vacuum, high temperature, cryogenic conditions, radiation.  Development and optimization of functional coatings: Formulation of fluorine-free coatings to reduce friction and wear and to protect against contamination and adhesion.  Analysis of electrical contact resistances, service life behavior, efficiency or stiffness properties in tribological systems.  European cooperation and technology transfer: AAC as an experienced partner in the development of innovative materials and surface solutions and in strengthening Europe’s technological sovereignty. Phone: +43 2622 90550-0 | E-mail: office@aac-research.at | Web: https: / / www.aac-research.at/ ger/ i-TRIBOMAT GmbH is the ÖTG’s most recent start-up, formed in 2022. The company sees itself as a Europe-wide point of contact for support regarding friction, wear and lubrication, offering services in the field of tribology and data-based services:  Tribological tests on more than 130 test benches and accompanying analyses with a wide range of test conditions, such as high and low temperatures, vacuum, aggressive environments (e.g., H 2 , NH 3 ) at the model contact and component level (e.g., bearings, coatings, gears, seals).  TRIBOLOGY UNIVERSE TM for modelling and simulation.  Access to the world’s largest database of materials and their tribological properties.  Customized databases for R&D laboratories.  Consulting on R&D activities in tribology. Phone: +43 676 9257923 | E-mail: office@i-tribomat.eu | Web: https: / / i-tribomat.eu/ de/ News 84 Tribologie + Schmierungstechnik · volume 72 · issue 3-4/ 2025 Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikationswiss chaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwiss chaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Tourismus \ VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ Altphilol Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikatio issenschaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Spra issenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Tourismus \ VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ hilologie \ Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwese remdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwissenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Touris VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ Altphilologie \ Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwiss chaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikationswissenschaft \ Linguistik \ Literaturgeschichte \ Anglisti auwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwissenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtsc BUCHTIPP Nicole Dörr, Carsten Gachot, Max Marian, Katharina Völkel 24th International Colloquium Tribology Industrial and Automotive Lubrication Conference Proceedings 2024 1. Auflage 2024, 279 Seiten €[D] 148,00 ISBN 978-3-381-11831-1 eISBN 978-3-381-11832-8 expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 \ 72070 Tübingen \ Germany Tel. +49 (0)7071 97 97 0 \ Fax +49 (0)7071 97 97 11 \ 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 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. 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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 Merle Reimers, Silvia Richter, Georg Jacobs, Joachim Mayer, Florian König Influence of Corrosion Inhibitors on the Wear Protection of Extreme Pressure / Anti Wear Additives in Oil-lubricated Rolling Bearings Nadja Aufderstroth, Lennart Schierholz, Jaacob Vorgerd, Manuel Oehler Test method using non-circular discs with a locally varying slide-to-roll ratio to investigate scuffing Christian Spura Wear energy density for wear prediction of displaceable spline couplings Felix Bernhardt, Katrin Alt, Markus Wöppermann Systematic investigation of the µ-mechanical material change of the sealing edge of radial shaft seals Wilhelm Rehbein, Isabell Lange, Kevin DiNicola, Salvatore Rea, John Williams EP Additives with Enhanced Sustainability for Water-miscible Metalworking Fluids Markus Grebe, Henrik Buse, Richard Heinlein, Andreas Keller Modern Application-Oriented Tribometry: Understanding Tribology Instead of Producing Characteristic Values—How Modern Measurement Technology Can Enable a View into the Hidden Tribological Contact Katrin Alt, Markus Wöppermann, Frank Plenert, Jürgen Liebrecht Design and Implementation of a Model Test for Reproducing Thickener Degradation in Grease-Lubricated Gearboxes and Its Role in Grease Formulation Development Ricardo Crespo Martins, Dennis Konopka, Mareike Dukat, Florian Pape, Martin Nicolaus, Kai Möhwald, Gerhard Poll, Max Marian Molybdenum APS-Coatings: A Self-regenerative Solution for Wear Resistance in Dry-lubrication Applications Hoang Viet Le, Kai Weigel Influence of tool coatings on the tribological effectiveness of bio-lubricants for sheet metal forming Adrian Heinl, Erich Prem, Christian Wilbs, Fabian Kaiser, Daniel Frölich Optimization of Radial-Shaft-Sealing systems for the use at high circumferential speeds Dennis Jess, Andreas Ligocki Concept for the generation of single-type wear particles for training the AI-based image processing of a particle sensor