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Tribologie und Schmierungstechnik EDITOR IN CHIEF MANFRED JUNGK 6 _ 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 6 | 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. When no source is indicated, all rights to pictures are reserved by the author or the editorial office. No third-party claims can be made unless otherwise agreed upon. The editorial office retains the right to edit and shorten articles. Trade names and commercial names mentioned in this journal may not be readily used by everyone, as they are often registered and protected trademarks. The journal, including all articles and pictures, is protected by copyright law. Excluding legally permitted cases, further use of the content without the publisher’s consent is punishable by law. This applies especially to copying, translating, creating microfilms, and using and processing the content in electronic systems. All information in this journal has been compiled with great care. However, mistakes cannot be ruled out entirely. Therefore, neither the publisher nor the authors assume liability for the correctness of the content or any mistakes and their consequences. 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 978-3-381-13811-1 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 6/ 2025 DOI 10.24053/ TUS-2025-0028 There is a big difference amongst the German speaking population in Europe when it comes to wine patriotism. Unlike in Austria, Switzerland, or South Tyrol in Germany customers reach for the mass-produced Spanish or Italian wines because of far too little awareness of local quality. The share of German local wine consumption has dropped to 42 percent from 48 percent in 2013. The other reason why German viticulture is in the midst of a severe, for many winemakers even existential crisis, is that it has become fashionable to condemn wine as a neurotoxin and a life-destroyer; the World Health Organization even considers the smallest amounts of alcohol harmful and preaches complete abstinence. I remember well the comment of the German wine associations president 2 decades ago when he was faced with the statistics that per capita the annual wine consumption dropped to 26 liters, “what would he drink on January 27 th ”. In Germany, declining demand is not matched by a decreasing supply. Unlike France, Spain, and Italy, which have lost almost half their vineyards, the area under vine has remained constant at around 100,000 hectares for years. This stability is partly due to the fact that wine is not just another consumer good for its producers, which consist of many, for generations family-owned estates driven more by emotional than economic principles. There are other countries such as Japan, Canada, Great Britain, the Netherlands, and the United States, where wine consumption is increasing. However, Germany’s wine export has increased only 11 percent since 1990, while Chile’s have grown by 550 percent and New Zealand’s by a staggering 2,600 percent. Overall, global wine exports have doubled in the past 35 years, but Germany still plays a minor role and even ranks behind South Africa and Portugal in exports. The number of wineries in Germany marketing their own wine has declined by a fifth in the past ten years, according to estimates from the Geisenheim University of Applied Sciences. They predict that half of Germany’s winegrowers will have to give up within the next ten years, and a third of the total vineyard area can no longer be cultivated profitably. Geisenheim University’s research includes the entire value chain of specialty crops, including grapevines, fruits, vegetables, and ornamental plants. They have facilities (Free Air Carbon dioxide Enrichment) to enrich the carbon dioxide by 20 % to study the effects of the increased CO 2 content on vine physiology, berry substances and yield stability. The tribological aspects of wine were presented by Nick Spencer at the last GfT Conference. Wine consists of countless compounds, many of which have a characteristic taste or effect in the mouth - the so-called mouthfeel. Mouthfeel is primarily due to the interactions between wine and saliva, which affect friction and lubrication in the mouth and can be detected with tribometers. To answer the question from above: we don’t hope so! Drinking less but better wines which are produced locally have a better climate footprint. And even for wine one can find that Tribology is everywhere. Your editor-in-chief Manfred Jungk Cheap is the way to the end? Events 2 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 Events Date Place Event ► 17.05. - 21.05.26 New Orleans, Louisiana (USA) 80 th STLE Annual Meeting & Exhibition ► 03.06. - 04.06.26 Bilbao, Spain 10 th Lubmat Conference ► 08.06. - 11.06.26 Palm Springs, CA (USA) NLGI Annual Meeting ► 02.09. - 04.09.26 Valpre, France 51 th Leeds Lyon Symposium ► 08.09. - 11.09.26 Gdansk, Poland 44 th Polish Tribology Conference ► 15.09. - 17.09.26 Düsseldorf, Germany Lubricant Expo Europe ► 20.09. - 25.09.26 Rio de Janeiro, Brazil 8 th World Tribology Congress ► 28.09. - 30.09.26 Wernigerode, Germany 67 th Annual GfT Conference ► 30.09. - 02.10.26 Amsterdam, The Netherlands 61 st UEIL Annual Congress 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 Contents 3 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 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 6 May 2026 5 Timo Schmidt, Oliver Methner, Christian Brecher, Dieter Mevissen, René Greschert Repeatability of the A10/ 16.6R/ 90 scuffing test for e-mobility oils 14 Shashivar Syla, Oliver Koch, Moritz Keuthen, Ralf Wuthenow A Component Independent Friction Model for Machine Elements 25 Dirk-Olaf Leimann Can lubricants with lower frictional torque in rolling contacts of bearings, gears and machines significantly reduce global warming? 34 Martin Strangfeld, Susanne Fritz Practical noise or stick-slip prevention in automotive interiors 1 Editorial Cheap is the way to the end? 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 6/ 2025 MEDIENTIPP Erik Kuhn On the Tribology of Lubricating Greases An energetic approach to post-modern tribology Tribologie - Schmierung, Reibung, Verschleiß 1. Auflage 2025, 204 Seiten ISBN print 978-3-381-14171-5 ISBN eBook 978-3-381-14172-2 DOI 10.24053/ 9783381141722 Ladenpreis print €[D] 118,00 Ladenpreis eBook €[D] 94,99 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 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 Introduction The ability of a lubricant to prevent scuffing damage in a gear transmission can be tested according to test methods A/ 8.3/ 90 [DIN25a] and A10/ 16.6R/ 90 [DIN25b]. These test methods and their derivatives have become elements of approval processes for gear lubricants in many applications [DENN22, DREC23, BART24]. Method A10/ 16.6R/ 90 is mostly used for lubricants in the automotive transmission sector. The test result is the load stage at which scuffing damage first occurs, for example FLS = 10 (Failure Load Stage). A scuffing failure is characterized by streaky roughening of the tooth flank surface in the direction of the profile, see Figure 1. DIN ISO 14635-2 [DIN25b] specifies the repeatability of the test with r = 1 load stages. However, it is known from literature that this value is not always achieved, especially for oils from the e-mobility sector [KADA21]. For this reason, repeat scuffing tests are often performed instead of one test per oil, which increases development costs. Incomplete running-in processes in the first test intervals are discussed as possible reasons for the dispersion of test results [KADA21]. Approaches to reducing the dispersion of results include changing the direction of rotation [KADA21] or increasing the rotational speed [PELL22, DREC23]. The test gears have not been subject to variation and optimization so far. Science and Research 5 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0029 Repeatability of the A10/ 16.6R/ 90 scuffing test for e-mobility oils Timo Schmidt, Oliver Methner, Christian Brecher, Dieter Mevissen, René Greschert* submitted: 13.10.2025 accepted: 06.01.2026 Presented at GfT Conference 2025 The A10/ 16.6R/ 90 scuffing test is an established test method to assess the scuffing load capacity of transmission lubricants. DIN ISO 14635-2 specifies a repeatability of r = 1 load stage for the test. However, this value is not always applicable to e-mobility gear oils, especially in case of low-viscosity fluids. An investigation is carried out to see if repeatability can be improved by using other test gear grinding processes and different gear flank roughnesses. Keywords Gear, Manufacturing, Grinding, Scuffing, Standardization, Gear Oil Abstract * Timo Schmidt, Oliver Methner Mercedes-Benz AG, 70327 Stuttgart Prof. Dr.-Ing. Christian Brecher Orcid-ID: https: / / orcid.org/ 0000-0002-8049-3364 Dr.-Ing. Dieter Mevissen Orcid-ID: https: / / orcid.org/ 0000-0002-9369-6363 Dr.-Ing. René Greschert (corresponding author) Orcid-ID: https: / / orcid.org/ 0000-0003-4167-8515 Laboratory for Machine Tools and Production Engineering (WZL) of RWTH Aachen University, 52074 Aachen MAAG grinding profile grinding Figure 1: Scuffing damages on Maag-ground and profile-ground A10 type gear flanks has fallen out of industrial use since the 1980s - except for the manufacture of A10 type gears. Test Gears This study investigated how the production-induced variation in the properties of type A10 test gears (Figure 1, Figure 2) could be considered a cause for the dispersion of test results in the scuffing test. As part of the study, in addition to the Maag grinding specified in DIN ISO 14635-2, the manufacturing processes of profile grinding and superfinishing were used to produce test gears. Furthermore, the surface roughness of the tooth flanks is part of the variation. DIN ISO 14635-2 specifies an average tooth flank roughness of Ra = 0.35 ± 0.1 µm for the test gears when the Ra values are measured in the lead direction [DIN25b]. However, if the roughness is measured on the same gears in the profile direction, i.e. in the direction of the friction in the test as specified for example in VDI guideline 2612-5 [VDI15], then values in the range of Ra = 0.6 ± 0.2 µm result for these gears. Since the roughness values in actual e-mobility gears are significantly lower than Ra = 0.6 µm, test gears with Ra = 0.25 µm and Ra = 0.10 µm were also included in the study. Apart from the Ra mean value, Maag grinding also shows different profile geometry characteristics than modern grinding methods such as profile grinding in terms of the dispersion of Ra between neighboring teeth of the same gear wheel or between different measuring sections on the same tooth, see Figure 3. Unlike profile grinding, Maag grinding does not require the use of metalworking fluid [KLOC24]. This makes it particularly attractive for the production of test gears for the investigation of gear lubricants. However, this advantage also means that the surface Ra in Maag grinding Science and Research 6 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0029 State of the Art Scuffing is a damage mechanism on tooth flanks that usually occurs spontaneously due to excessive loading or insufficient lubricating film, so that the contacting tooth flank surfaces are no longer separated. Adhesion occurs due to complex mechanical, thermal, and chemical interactions between the metal surfaces and lubricant components such as EP additives [MICH87, CEC18, KADA21, KLOC24]. Due to this complexity, there are no calculation methods for scuffing damage that are as reliable as those for pitting or tooth root damage, which is why experimental scuffing tests are still required today to qualify the scuffing load-carrying capacities of new gear oils. A widely used scuffing test in the automotive sector is the A10/ 16.6R/ 90 test, in which a pair of test gears with A10 type geometry is loaded in a back-to-back gear test rig and subjected to a step-bystep load stage increase [DIN25a, DIN25b]. The load stage at which scuffing damage first occurs is defined as the Failure Load Stage (FLS). The corresponding standard ISO 14635-2 requires that the test gears must be finished using Maag 15° criss-cross grinding, a grinding process developed in the 1910s and based on the kinematics of discontinuous generating grinding [KLOC24]. This grinding process works without metalworking fluids and creates a cross pattern on the tooth flanks (Figure 1), which facilitates photographic documentation during the scuffing test. The disadvantages of Maag grinding compared to modern methods such as profile grinding or continuous gear grinding are significantly longer machining time per workpiece due to dry machining, and result in less reproducible quality and surface roughness. For example, the Ra values of Maag-ground A10 gears can vary from tooth to tooth between Ra = 0.5 µm and Ra = 0.9 µm (see Figure 3) when the roughness measurement is conducted in the profile direction, i.e., parallel to the force acting in the scuffing test [VDI15]. For this reason, Maag grinding Pinion Wheel ➢ Gear grinding process − Maag grinding (standard) − profile grinding − superfinishing ➢ Gear flank rougness − Ra = 0.60 μm (standard) − Ra = 0.25 μm − Ra = 0.10 μm ➢ Micro geometry − no modification (standard) − crowning C β 1/ 2 = 0.5 / 4.0 μm alongside contact width Geometry Number of teeth z 1/ 2 = 16/ 24 Module m n = 4.5 mm Pressure angle α n = 20 ° Helix angle β = 0 ° Tooth width b 1/ 2 = 10/ 20 mm Profile shift x n1/ 2 = 0.8532 / -0.5 Pinion tip diameter d a1 = 88.77 mm Wheel tip diameterd a2 = 112.5 mm Quality class IT = Q5 Manufacturing Material 16MnCr5 (1.7131) Heat treatment Case hardening Surface hardness 60 - 62 HRC CHD 550HV 0,6 - 0,9 mm Shot peening of the wheel Variation A10-mod Specification A10 Figure 2: Test gear geometry A10 type and variation in this study (A10-mod) cannot be further reduced without the risk of thermal damage to the microstructure. It seems possible that the use of Maag-ground gears could sometimes result in more or less favorable combinations of tooth pairs and surface roughness of specific teeth. This effect is amplified by the fact that the numbers of teeth 16 and 24 are identical and that it is often observed that, especially in early failures in the scuffing test, only one tooth group and not all tooth groups show scuffing damage. Test Rig The investigations regarding the scuffing load capacity were carried out on a back-to-back gear test rig with a center distance of a = 91.5 mm, see Figure 4. The power circuit comprised a test gearbox, a reference gearbox, a torsional shaft, and a clutch. The gear set of the reference gearbox was designed to be significantly wider than that of the test gearbox so that damage occurred exclusively Science and Research 7 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0029 Maag grinding (A10) Profile grinding (A10-mod) standard deviation s(Ra) = 0.060 μm standard deviation s(Ra) = 0.012 μm Figure 3: Tooth-to-tooth variation of Ra for A10 type test gear wheel z 2 = 24 2 3 4 5 6 1 ◼ Key 1 Motor 2 Reference gearbox 3 Torsional shaft 4 Locating pin and clutch for torque application 5 Test gear set 6 Heating / cooling Torque cycle ◼ Center distance a = 91.5 mm ◼ Application of torque Lever + weight Figure 4: Back-to-back gear test rig and torque application Maag-ground gears in full accordance with DIN ISO 14635-2, Figure 5. The failure load stage (FLS) was achieved when the total area of damage to the 16 pinion teeth exceeded 100 mm 2 . During the experiments, the area of damage was examined using a stencil with a scale. One fully scuffed tooth flank corresponded to 70 mm 2 . For profile grinding, the four results of each class are each within the repeatability of r = 1. For the Maagground test gears, there is one FLS = 6, one FLS = 7, and two early failures at FLS = 4. Since all test gears were examined in 100 % all-tooth measurements before and after the scuffing tests, it was determined that there were particularly large differences between the mean Ra values of the pinion and the wheel for the two early failures. According to the theory of incomplete running-in [KADA21], the difference between the starting Ra values of pinion and wheel could have had an unfavorable effect on the running-in process here. Generally, it is well known from the literature that lower Ra values lead to lower friction and wear in tooth flank contact. It has also been proven for damage mechanisms such as pitting, micropitting and scuffing that low Ra values of the tooth flanks increase the gear load capacity [NISK05, HÖHN10, OHNO20]. However, in the A10/ 16.6R/ 90 scuffing tests of our study, the Ra = 0.25 µm gear sets achieved lower FLS than the Ra = 0.60 µm gear sets. It seems possible that, at least for the A10/ 16.6R/ 90 test with the low-viscosity lubricant used in this study, higher Ra values of the test gears were advantageous for some reason. One explanation could be plastic deformation additives in the oil that are particularly effective on rough surfaces [LOHN17]. Since the profile-ground gear sets with a starting roughness of Ra = 0.25 µm showed the lowest dispersion Science and Research 8 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0029 on the gears of the test gearbox. The test torque was applied by means of a weight load. To do this, one half of the open clutch was locked with a pin, while a lever was attached to the other half in a form-fitting manner and loaded with weights. This led to a relative rotation of the clutch halves, which was held in place by friction when the clutch was closed in the power circuit, maintaining a tensioning torque. This principle corresponds to the standardized test setup according to DIN ISO 14635-1 [DIN25a]. The test gears were lubricated by splash lubrication up to the axle center. After each test, the oil was discarded and replaced with fresh oil of the same type for the next test. The rotational speed of the gear wheel, which was located on the motor shaft, was set to n 2 = 2910 min -1 , while the pinion torque was stepwise increased up to T 1 = 373 Nm which reflected FLS = 10 as the highest load stage of the test. Before and during the study, the test rig function was checked by measuring the torque using strain gauges. Before the start of the study, the reference oils RL-214 and RL-215 of the CEC (Coordinating European Council) were tested on the test rig for their scuffing load capacities. The FLS values obtained were within the target ranges determined by other test laboratories in round robin tests for the CEC [CEC18]. Scuffing Tests A10/ 16.6R/ 90 scuffing tests were performed with the manufactured A10 test gears using BEV oil with a viscosity of KV40 < 20.5 cSt. Four scuffing tests were performed with each of the A10-mod test gears with the two Ra classes Ra = 0.25 µm and Ra = 0.60 µm. As a reference variant, four scuffing tests were performed with Profile grinding Ra=0.23-0.31 μm / Ra=0.53-0.62 μm Maag grinding acc. to ISO 14635-2 Ra=0.35-0.64 μm 0 1 2 3 4 5 6 7 8 Ra0.25 - Test 1 Ra0.25 - Test 2 Ra0.25 - Test 3 Ra0.25 - Test 4 Ra0.60 - Test 1 Ra0.60 - Test 2 Ra0.60 - Test 3 Ra0.60 - Test 4 Maag - Test 1 Maag - Test 2 Maag - Test 3 Maag - Test 4 Failure load stage in A10/ 16.6R/ 90 test ◼ Test gear A10-mod a = 91.5 mm z 1/ 2 = 16/ 24 m n = 4.5 mm α n = 20 ° β = 0 ° b 1/ 2 = 10/ 20 mm Batch 1 ◼ Test conditions 16.6R/ 90 n 2 = -2910 min -1 T Öl = 90 ° C BEV-Oil KV 40 < 20.5 cSt ◼ Failure criteria Scuffing surface >100 mm 2 Pinion Ra: Wheel Ra: 0.23 0.31 0.26 0.30 0.29 0.25 0.30 0.29 0.61 0.55 0.62 0.53 0.61 0.56 0.59 0.54 0.35 0.41 0.59 0.55 0.39 0.62 0.64 0.40 Figure 5: Influence of gear flank grinding on scuffing test of the FLS achieved in the 4 single scuffing test runs, further gear sets were manufactured using this roughness variant. The gear sets were manufactured using the same profile grinding machine setup but in different months in order to investigate the influence of production dispersion on the test procedure. Four tests were performed for each batch, see Figure 6. The FLS was FLS = 5 in eight tests. In the other four tests, FLS = 6 and FLS = 7 were achieved. It was therefore possible to determine an influence of the production batch on the FLS in the scuffing test, since it was mainly the batch 2 gears that obtained FLS = 6. However, all test parts met the specification of Ra = 0.25 µm ± 0.05 µm and the dispersion of FLS was still below the dispersion known from the standard test of that lubricant using Maag gears. To further characterize the influence of the initial Ra value of the tooth flanks on the results of the scuffing test, gear sets were subjected to additional superfinishing treatments. Two suppliers with different superfinishing processes were selected, A and B. In both cases, these processes reduced the tooth flank roughness from Ra = 0.25 µm to Ra = 0.10 µm. The superfinished gear sets were then tested in scuffing tests using the above parameter set and BEV oil; see Figure 7 for the results. It has been reported in the literature that superfinishing can lead to lower friction and wear in tooth flank contact [WINK08, NISK13]. However, superfinishing did not lead to an increase of the achievable FLS in the A10/ 16.6R/ 90 scuffing tests of this study. The average FLS for both superfinishing variants was FLS = 5 Science and Research 9 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0029 0 1 2 3 4 5 6 7 8 Batch 1 - Test 1 Batch 1 - Test 2 Batch 1 - Test 3 Batch 1 - Test 4 Batch 2 - Test 1 Batch 2 - Test 2 Batch 2 - Test 3 Batch 2 - Test 4 Batch 3 - Test 1 Batch 3 - Test 2 Batch 3 - Test 3 Batch 3 - Test 4 Failure load stage in A10/ 16.6R/ 90 test ◼ Test gear A10-mod a = 91.5 mm z 1/ 2 = 16/ 24 m n = 4.5 mm α n = 20 ° β = 0 ° b 1/ 2 = 10/ 20 mm Batch 1-3 ◼ Test conditions 16.6R/ 90 n 2 = -2910 min -1 T Öl = 90 ° C BEV-Oil KV 40 < 20.5 cSt ◼ Failure criteria Scuffing surface >100 mm 2 Figure 6: Influence of grinding batch (month of production) on scuffing test ◼ Test gear A10-mod a = 91.5 mm z 1/ 2 = 16/ 24 m n = 4.5 mm α n = 20 ° β = 0 ° b 1/ 2 = 10/ 20 mm Batch 1 ◼ Test conditions 16.6R/ 90 n 2 = -2910 min -1 T Öl = 90 ° C BEV-Oil KV 40 < 20.5 cSt ◼ Failure criteria Scuffing surface >100 mm 2 0 1 2 3 4 5 6 7 8 Ra0.25 - Test 1 Ra0.25 - Test 2 Ra0.25 - Test 3 Ra0.25 - Test 4 Finish A - Test 1 Finish A - Test 2 Finish A - Test 3 Finish B - Test 1 Finish B - Test 2 Finish B - Test 3 Finish B - Test 4 Failure load stage in A10/ 16.6R/ 90 test Figure 7: Influence of superfinishing on scuffing test regarding failure load stage The loss-of-lubrication test setup of M ORHARD might be more similar to the A10/ 16.6R/ 90 test of low-viscosity oils than the setups of the other studies mentioned above. An explanation for the observed effect is suspected to be oil retention mechanisms in the microtexture [MORH23]. The results of our study suggest that scuffing damage occurs later when test gears with higher starting Ra values are used. This positive correlation between Ra and FLS contradicts the state of the art, according to which friction and thus also the risk of scuffing increase with high surface Ra values of the gears [MICH87, HÖHN10, CZIC20, VORG23, HONG25]. Nevertheless, the effect was reproducible in additional tests with profile-ground test gears with very high Ra (see Figure 9). In regard to the significantly different appearances of the tooth flanks with and without superfinishing after the scuffing tests, the FLS = 5+ and FLS = 5were graphically represented in the diagram by increased and reduced markings, respectively. Since experience from industrial applications shows that the scuffing load capacity is increased, and not decreased, by low Ra surfaces, the effect observed here could be an artifact in this study’s test setup, and especially in the A10/ 16.6R/ 90 test of the investigated BEV-oil. It is possible that the running-in effect of the test gears is so decisive for this test method that the test is influenced more by improvements in the running-in processes, e.g. through plastic deformation additives [LOHN17], than by anti-scuffing additives. It therefore remains to be investigated whether the A10/ 16.6R/ 90 test still reflects real-world application for e-mobility oils and whether oils are already equipped with additives for surface smoothing in order to pass the current approval process, although these additives are not needed in real-world application. Science and Research 10 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0029 which was the same level as the non-superfinished gears achieved. During and after the scuffing tests, the appearance of the tooth flanks was documented by photos, see Figure 8. Since the teeth with scuffing damage looked identical for all variants and did not allow any differentiation, photos of the best-case tooth flanks were taken instead. These were the tooth flanks that showed the least wear and scuffing in the respective tests after the respective load stages. For the standard variant with Ra = 0.25 µm, after many tests, teeth were found that showed almost no wear marks, with the exception of a few scratches. However, on the gears of the superfinished variants, all teeth were affected by deep grooves and pronounced scuffing zones. For this reason, the tests of the non-superfinished variant were rated “FLS = 5+” and the tests of the superfinished variants were rated “FLS = 5-”. It is remarkable that the low FLS in the scuffing tests of superfinished gear sets contradict findings from industrial practice and the current state of the art regarding the influence of grinding processes and superfinishing on scuffing load capacity. M ICHAELIS for Maag-ground gears, H ÖHN for vibratory ground gears, H ONGU for honed gears and V ORGERD for isotropically superfinished gears have observed that the scuffing load capacity of gears tends to increase when the test gears show low starting Ra values. However, the majority of those studies were performed with test setups other than A10/ 16.6R/ 90 and with oils of higher viscosities [MICH87, HÖHN10, VORG23, HONG25]. The study of M ORHARD on the scuffing load capacity of superfinished gears under loss-of-lubrication conditions (sudden dry running after continuous injection lubrication of FVA3A oil) in an A/ 8.3/ 90 test setup reveals that superfinished gears are subject to scuffing damage more quickly than conventionally ground gears [MORH23]. LS4 LS5 Not superfinished Ra = 0.25 μm Superfinished variant A Ra = 0.10 μm Superfinished variant B Ra = 0.10 μm „FLS = 5+“ „FLS = 5 - “ „FLS = 5 - “ Figure 8: Unscuffed gear flanks after A10/ 16.6R/ 90 scuffing tests of superfinished gears Summary and Outlook DIN ISO 14635-2 specifies a repeatability of r = 1 load stage for the A10/ 16.6R/ 90 scuffing test. This value is not always applicable for gear oils from e-mobility applications. Therefore, test gears were manufactured using more modern grinding processes and narrower specifications for tooth flank Ra in this study. With the modified test gears, better repeatability of the FLS in the scuffing test was achieved for the lubricant considered in this study. Furthermore, a positive correlation between the start Ra of the test gears and the FLS in the scuffing test was observed in the investigations. Against this background in particular, it remains to be examined whether the current test gear geometry of the A10/ 16.6R/ 90 scuffing tests reflects the conditions of transmissions in real e-mobility applications. Literature [BART24] Barth, Y. J.; Sagraloff, N.; Egger, G.; Tobie, T.; Stahl, K.: Investigations on Ways to Improve the Scuffing and Wear Behavior of Oil-Free Water- Based Lubricants for Gear Applications. In: ASME Journal of Tribology, Vol. 146, No. 5, 2024, doi: 10.1115/ 1.4064401 [CEC18] Test method description L-84-02: FZG Scuffing Load Carrying Capacity Test for High EP Oils. Coordinating European Council, 2018 [CZIC20] Czichos, H.; Habig, K.-H.: Tribologie-Handbuch - Tribometrie, Tribomaterialien, Tribotechnik. 5. Aufl., Springer Vieweg, Wiesbaden, 2020 [DENN22] Dennig, H.-J.; Zumofen, L.; Stierli, D.; Kirchheim, A.; Winterberg, S.: Increasing the Safety against Scuffing of Additive Manufactured Gear Wheels by Internal Cooling Channels. In: Forschung im Ingenieurwesen, Vol. 86, No. 4, 2022, pp. 595-604, doi: 10.1007/ s10010-021-00515-5 [DIN25a] DIN ISO 14635-1: Gears - FZG Test Procedures - Part 1: FZG Test Method A/ 8,3/ 90 for Relative Scuffing Load-Carrying Capacity of Oils. Beuth, Berlin, 2025 [DIN25b] DIN ISO 14635-2: Gears - FZG Test Procedures - Part 2: FZG Step Load Test A10/ 16,6 R/ 120 for Relative Scuffing Load-Carrying Capacity of High EP Oils. Beuth, Berlin, 2025 [DREC23] Drechsel, A.; Pellkofer, J.; Stahl, K.: Transferability of the Scuffing Load Capacity of Gear Oils Determined on Spur Gears to Hypoid Gears. In: Forschung im Ingenieurwesen, Vol. 87, No. 5, 2023, pp. 923-932, doi: 10.1007/ s10010-023-00686-3 [HÖHN10] Höhn, B.-R.; Tobie, T.; Koller, P.: Steigerung der Zahnflankentragfähigkeit durch Kombination von Strahlbehandlung und Finishingprozess. FVA-Forschungsvorhaben Nr. 521 I, Abschlussbericht, FVA-Heft Nr. 957, Frankfurt am Main, 2010 [HONG25] Hongu, J.; Hashimoto, S.; Kato, J.; Inawaka, T.; Koide, T.: Scuffing Properties of Gear Pairs with Different Surface Textures (Grinding, Small- Angle Honing, and Barreling). In: Forschung im Ingenieurwesen, Vol. 89, No. 2, 2025, doi: 10.1007/ s10010-025-00862-7 [KADA21] Kadach, D.; Michaelis, K.; Hein, M.; Tobie, T.; Stahl, K.: Fresstragfähigkeit von Schmierstoffen für Doppelkupplungsgetriebe. In: antriebstechnik, Vol. 60, No. 1, 2021, pp. 60-66 [KLOC24] Klocke, F.; Brecher, C.: Zahnrad- und Getriebetechnik: Auslegung - Herstellung - Untersuchung - Simulation. 2. Aufl., Hanser, München, 2023 [LOHN17] Lohner, T.; Mayer, J.; Michaelis, K.; Höhn, B.-R.; Stahl, K.: On the Running-in Behavior of Lubricated Line Contacts. In: Proceedings of the Institution of Mechanical Engineers, Part J: Journal of Engineering Tribology, Vol. 231, No. 4, 2017, pp. 441-452, doi: 10.1177/ 1350650115574869 [MICH87] Michaelis, K.: Die Integraltemperatur zur Beurteilung der Fresstragfähigkeit von Stirnradgetrieben. Dissertation, Technische Universität München, 1987 Science and Research 11 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0029 4 5 6 7 8 9 10 0 0,1 0,2 0,3 0,4 0,5 0,6 0,7 0,8 0,9 1 Failure load stage in A10/ 16.6R/ 90 test ◼ Test gear A10-mod a = 91.5 mm z 1/ 2 = 16/ 24 m n = 4.5 mm α n = 20 ° β = 0 ° b 1/ 2 = 10/ 20 mm ◼ Test conditions 16.6R/ 90 n 2 = -2910 min -1 T Öl = 90 ° C ◼ Failure criteria Scuffing surface >100 mm 2 gear set Ra / μm CEC Reference Oil 215 KV 40 = 64.6 cSt C β 1/ 2 = 0 / 0 μm BEV-Oil KV 40 < 20.5 cSt C β 1/ 2 = 0.5 / 4.0 μm (this study) Positive correlation between Ra and FLS 5+ 5all gear sets profile-ground Figure 9: Influence of test gear Ra on FLS [PELL22] Pellkofer, J.: Method to Assess the Scuffing Load Capacity of Lubricants for Gears in E-Vehicles Using an FZG Gear Test Rig. In: STLE E-Mobility Conference, San Antonio, 30.11.-02.12.2022 [VDI15] VDI/ VDE 2612-5: Measurement and Testing of Gearings - Surface Roughness Measurement of Cylindrical Gears and Bevel Gears by Means of Stylus-Type Instruments. VDI, Berlin, 2015 [VORG23] Vorgerd, J.; Tenberge, P.; Steinrötter, M.: Scuffing Load-Carrying Capacity of High-Speed Gears with an Isotropic Superfinished Surface. In: International Conference on Gears 2023, VDI-Berichte Nr. 2422, VDI Verlag, Düsseldorf, 2023, pp. 227-242, EID: 2-s2.0-85175043442 [WINK08] Winkelmann, Lane; El-Saeed, Omer; Bell, Matt.: The effect of superfinishing on gear micropitting, part II. REM Chemicals, Inc.. In: AGMA Fall Technical Meeting, 2008 Science and Research 12 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0029 [MORH23] Morhard, B.; Lohner, T.; Stahl, K.: Influence of Surface and Material Technologies on Loss of Lubrication Performance of Gears. In: VDI Berichte Nr. 2422, VDI Verlag, Düsseldorf, 2023, S. 779- 784 [NISK05] Niskanen, P., Hansen, B., Winkelmann, L.: Evaluation of the Scuffing Resistance of Isotropic Superfinished Precision Gears. AGMA Technical Paper 05FTM13, 2005 [NISK13] Niskanen, P.; Manesh, A.: Reducing wear with superfinish technology. The AMPTIAC Quarterly, Volume 7, Number 1, 2003 [OHNO20] Ohno, T.; Shiota, T.; Fujii, M.: The influence of lubricant additives and surface roughness and hardness of material on the damage behavior of gears. Tribology International, 143, Article 106492, 2020 Science and Research 13 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 Attend our seminars, courses and conferences. Friction, wear and lubrication Lubricants and operating fluids Lubrication technology Lubricated machine elements A large part of our seminars is supported by the Ministry of Economic Affairs, Labour, and Housing of Baden-Württemberg with funds from the European Social Fund. Benefit from the ESF course funding and secure up to a 70 % subsidy on your participation fee. All information on eligibility for funding can be found at www.tae.de/ foerdermoeglichkeiten Tribology, friction, wear and lubrication Up to 70 % subsidy possible Further information and registration at www.tae.de/ weiterbildung/ tribologie-reibung-verschleiss-schmierung/ Science and Research 14 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 1 Introduction The design of modern drivetrains and individual machine elements relies heavily on contact simulations. Within the Research Association for Drive Technology (FVA), numerous contact and friction models have been developed over the years. These models are usually semiempirical and have evolved over time. They are often specific to machine elements. As a result, it is difficult to compare research findings, as different calculation tools often lead to significantly different results. In engineering practice, this necessitates the use of high safety factors. The first fundamental approaches to friction modeling were established by Sjovall [1], Stribeck [2], Lundberg and Palmgren [3-5]. Their pioneering work laid the foundation for understanding tribological losses but was limited to simplified relationships. With the advent of more complex computer-based models, the representation of rolling and sliding contacts was gradually refined. Today, the universal friction model enables a high-resolution and significantly more accurate description of contact mechanics. The universal friction model is optimized for the calculation of general discretized contacts and allows for holistic tribological system optimization of gear transmissions. It provides the basis for reliable calculation of the CO 2 -equivalent (Co 2 e) footprint during the lifetime of the gearbox, enables the consideration of electrical properties of transmissions systems, and serves as a key enabler for advanced bearing and gear design. A central aspect is the precise determination of frictional losses, which play a critical role in drivetrain efficiency, system performance, and overall energy consumption [6-8]. These frictional losses predominantly occur in highly loaded rolling and sliding contacts of gears, bearings, and other mechanical components [9]. In such concentrated contacts, pressures of several giga- A Component Independent Friction Model for Machine Elements Shashivar Syla, Oliver Koch, Moritz Keuthen, Ralf Wuthenow * submitted: 25.09.2025 accepted: 03.12.2025 Presented at GfT Conference 2025 The calculation of friction losses is a key element in optimizing modern drivetrain systems. However, many simulation environments still rely on semi empirical approaches, which are often developed separately for individual machine elements such as bearings or gears. This fragmentation frequently leads to inconsistencies and hampers comprehensive systemlevel analysis and optimization. A component-independent friction model addresses these challenges with a physically based approach that requires no element-specific adaptations. It is based only on the geometric and material properties of the contact bodies, making it independent of specific machine elements. The model can be used for both rolling bearings (e.g. deep groove ball bearings, angular contact ball bearings, cylindrical roller bearings) and gears such as cylindrical gears, without changing the underlying friction formulation. This contribution focuses on the application to rolling bearings, where the friction torque is resolved locally and calculated with high physical accuracy. Local contact parameters such as contact pressure, lubricant film thickness, and relative velocity are considered at each rolling contact. In addition to load-dependent friction losses, the model also captures load-independent effects, including seal friction, hydraulic losses, and the load-free zone. The model serves as the new standard friction model in the FVA-Workbench, further underlining its relevance and robustness for practical applications. Integration into the FVA-Workbench makes it possible to change the internal geometry of a bearing in order to carry out detailed studies on design optimization and its influence on friction behavior, as well as to take into account the complex interactions in the mechanical system of the entire drive train. Model outputs have been compared to experimental data to validate the results. The component-independent modeling approach supports the analysis of friction losses and contributes to the efficient and robust design of drivetrain systems. Keywords tribology, friction, friction torque, rolling bearing, simulation, experimental validation Abstract * Shashivar Syla, M.Sc. 1 (corresponding author) Prof. Dr.-Ing. Oliver Koch 1 Dr. rer. nat. Moritz Keuthen 2 Dr.-Ing. Ralf Wuthenow 2 1 RPTU Kaiserslautern-Landau, Chair of Machine Elements, Gears and Tribology, Gottlieb-Daimler Str. 42, D-67661 Kaiserslautern, Germany. 2 FVA GmbH, Lyoner Str.18, D-660528 Frankfurt/ Main, Germany. pascals may arise [10], making reliable lubrication indispensable. Modeling these contacts is therefore the subject of numerous research projects in which specialized research software has been developed. For gears, examples include RIKOR [11-13], STIRAK [14-16], and SNETRA [17-19], while for bearings, the research software LAGER2 [20-22], is available. Some of these programs are accessible as research software, whereas others are offered exclusively as commercial products or remain proprietary company knowledge. For practical drivetrain design, quasi-static calculation tools such as Bearinx [23,24] (developed by Schaeffler) and the FVA- Workbench [25,26] are widely used. The universal contact friction model is a novel, unified methodology for simulating tribological contacts that is now available for use. Its integration into the FVA-Workbench for both gears and bearings establishes a common framework that not only enables a more reliable optimization of frictional losses but also ensures cross-element comparability - laying the foundation for advanced and sustainable drivetrain design. 2 Material and Methods As outlined in Section 1, predicting frictional losses in gear transmissions and rolling bearings requires an approach to calculation that is both generally applicable and sufficiently accurate. To achieve this, a universal friction model has been developed in this paper. This model can be used for both point and line contacts, enabling friction losses in gears and rolling bearings to be calculated within the same modelling framework. The theoretical background of the model is described in Section 2.3.1. Section 3 presents the experimental setup, Section 4 provides the results, and Section 5 contains the discussion. 2.1 Quasi-Static Friction Calculation The accurate design of gear systems requires consideration of all cross-effects between individual machine elements within the complete drivetrain under operating conditions. These interactions are governed by the elastic behavior of the components and result in complex deformation states. In quasi-static system-level calculations, both the power flow and the deformations of the entire gearbox are determined. The resulting forces and displacements are consistently transferred to the individual machine elements, enabling a realistic representation of the overall drivetrain behavior. Figure 1 illustrates the quasi-static friction calculation for rolling bearings and gears, which is composed of the three fundamental components: contact point determination, local contact calculation, and friction evaluation. This representation highlights the general modeling approach, in which component-specific interfaces are combined with the universal contact friction model. The component-specific interface is machine-element-specific. In the case of rolling bearings, it includes aspects such as kinematic iteration and determining the unloaded zone. For gears, it includes aspects such as calculating power loss along the line of action. The universal contact friction model, in contrast, provides a consistent description of friction by treating only the two contacting bodies, independent of the component type. The integration of component-specific formulations with the universal friction contact model, in combination with quasi-static system calculations, enables a coherent and physically consistent representation of drivetrain performance. 2.1.1 Contact Force Calculation Within the contact force calculation, a distinction is made between point and line contacts when calculating contact loads. Point contacts, as they occur in the raceway contacts of ball bearings and in the rib contacts of roller bearings, are calculated according to Hertz theory to determine the load-deformation relationship. Line Science and Research 15 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 Figure 1: Quasi-Static Friction Simulation rolling element are solved. For this purpose, the calculation model provides geometry, material properties, lubrication parameters, as well as the contact normal n R,i , penetration depth δ R,i , and pressure distribution p R,i for each roller and each of its contacts. Based on this information, the contact ellipses at the contact points are aligned. Within these contact ellipses, the friction forces are determined by calling the universal friction model for point and line contacts. The resulting forces are then transformed from the local ellipse coordinate system into the roller coordinate system, where they are used to iterate the kinematic solution. To implement the equilibrium, three degrees of freedom must be granted: • Cage can rotate around the bearing axis • Roller can rotate around its rolling axis • Roller can rotate around its bore axis The losses in the load-free zone amount to approximately 20 - 30 % of the total driving torque and are therefore not negligible. To calculate these losses, it is assumed that the mass forces cause the rolling element to be guided on the outer ring. The extent of the load-free zone depends on the load and must be determined on the basis of the load distribution. Since the load-free zone occurs exclusively in rolling bearings, only the universal contact friction model is considered in the following. Depending on the type of lubrication used, for example in oil bath lubrication, splashing losses are also taken into Science and Research 16 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 contacts, which predominantly occur in the raceway contacts of roller bearings and in the flank contacts of gears, are modeled using established slice approaches. In these models, the rolling element is discretized into slices along the rolling axis, with the stiffness of each slice derived from the total stiffness and distributed across the discretization points [27]. A limitation of conventional slice models is that the load applied to one slice does not influence the deformation of adjacent slices, preventing the accurate representation of edge stresses that may occur under misalignment or tilting conditions [27]. The advanced slice model (AST) overcomes this limitation by incorporating coupling effects between neighboring slices, which are necessary to capture such edge stresses [27]. 2.2 Component Specific Interface The friction torque of a rolling bearing M R is calculated using the component-specific interface. It is comprised of the contributions from the load zone M LZ , the unloaded zone M LFZ , the splash losses M P and the sealing losses M SL (see equation (1)). Figure 2 illustrates the sequence of this calculation process. (1) For the calculation of the friction torque from the load zone, an iterative determination of the kinematics is performed for all rolling elements. Both the force equilibrium in the circumferential direction and the moment equilibria around the rolling axis and the bore axis of the 𝑀 𝑅 = 𝑀 𝐿𝑍 + 𝑀 𝐿𝐹𝑍 + 𝑀 𝑃 + 𝑀 𝑆𝐿 Figure 2: Component Specific Interface account. Various calculation methods are available, including those proposed by Schaeffler [28], SKF [29], Timken [30] and ISO 14179-2 [31]. The determination of seal losses in sealed bearings is only possible using the SKF catalogue method. 2.3 Universal contact friction model The mathematical and physical foundations for describing friction at the contact points of rolling bearings are complex since different friction phenomena occur depending on the specific contact geometry and relative motion. The universal friction model accounts for all relevant factors that determine friction in the tribological contacts of rolling bearings and gears. Losses in rolling contacts can be divided into rolling friction losses and sliding friction losses. Rolling friction losses primarily arise from material hysteresis and lubricant compression in the inlet zone of the contact, whereas sliding friction losses result from differential slip, spin, and macroscopic sliding, for example in the rollerrib contact. [28] In the universal friction model, sliding friction losses are described as the sum of the shear of the lubricant, i.e. elastohydrodynamic lubricant friction (EHD), and sliding in solid-solid contact. To represent mixed friction conditions, these two components are weighted by the solid contact ratio. Rolling friction losses are taken into account by modelling lubricant compression in the runin area and material hysteresis within the lubricant friction. A central feature of the friction force calculation within the universal friction model is the iterative temperature adjustment. The temperature field is updated successively until a steady-state equilibrium is achieved. This procedure incorporates total power losses, which result from both lubricant friction and solid contact friction. Figure 3 provides an overview of the calculation process and illustrates the composition of the individual loss components. 2.3.1 Friction Components in EHL Contact In highly loaded, locally concentrated contacts, both rolling and sliding motions occur, which are opposed by friction forces. While rolling motion is mainly influenced by rolling friction, sliding motion is governed by sliding friction. The friction forces arising from fluid friction in the EHD contact result from shear gradients determined by the velocity distribution within the lubricant film [32]. In general, the flow can be divided into pressure flow and shear flow [32]. Pressure flow is particularly pronounced in the inlet region, where it leads to dominant rolling friction [32]. In contrast, its influence in the central contact area is negligible, so that shear flow - and thus sliding friction - prevails [32]. In the case of pure rolling motion, work must be expended to compress the lubricant in the inlet region, which is performed by the rolling friction force F roll [32]. The rolling friction force can be determined according to Crook [50] by evaluating the following integral over the entire contact area. Science and Research 17 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 Figure 3: Calculation process: Universal contact friction model Ψ ratio is used to measure the proportions of solid and fluid sliding friction. (7) In equation (7), F s represents the proportion of the normal contact force F n that is transmitted via the solid contacts [37]. The resulting sliding friction force F slide can be calculated using equation (8). (8) To visualize the sliding - and rolling friction forces, Figure 4 shows a contact loaded by the normal force and the frictional forces acting in the contact. Taking the existing coordinate system into account, the resulting frictional force between the contacting bodies is given by the relationship shown in equation (9). (9) The following variables are relevant for calculating the friction forces according to equation (9): Velocities in Contact, v sum , v rel Load F N Geometry of the bodies R x,y,1,2 , l 1,2 Material and Lubricant Properties η, ρ, E 1,2 , ν 1,2 In addition to the friction forces F roll,1/ 2 and F slide,1/ 2 , Figure 4 also shows the real pressure distribution p in yellow as well as the approximation of the pressure distribution according to Hertz as a dotted black line. The velocity field is considered as a linear distributed, being defined by ψ = 𝐹 𝐹 N 𝐹 = ψ ⋅ 𝐹 , + (1 − ψ) ⋅ 𝐹 , 𝐹 = −𝐹 ± 𝐹 Science and Research 18 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 (2) The contributions from the load zone and the outlet region to rolling resistance are minor and are therefore neglected in [33]. Another contribution to rolling resistance arises from material hysteresis [34]. During a load cycle, a portion of the applied energy is dissipated and is no longer available upon unloading [27]. The resulting resistance torque can be described for line contacts by the equation (3), according to Johnson [35]. (3) For point contacts, [35] provides the following expression: (4) The sliding friction force in the lubricant arises from the shearing of the fluid. Due to the relatively low pressures in the inlet region and the associated decrease in viscosity, shearing in this area can be neglected, so that the main contribution to sliding friction is determined by the high pressures within the contact [27]. The sliding component of fluid friction can be calculated using equation (5). (5) The solid component of sliding friction F slide,s can be calculated using the solid friction coefficient μ and the normal force F N [36]. (6) To model the mixed friction condition, the asperity load 𝐹 = 1 2 ⋅ ∫ ℎ 𝑥 𝑎 𝑥 𝑒 ⋅ ⋅ 𝑑𝑝 𝑑𝑥 𝑑x 𝑀 𝐻 = 𝑎 𝐻 ⋅ 𝐹 𝑁 ⋅ 2 ⋅ 𝑏 𝐻 3 ⋅ π 𝑀 𝐻 = 𝑎 𝐻 ⋅ 𝐹 𝑁 ⋅ 3 ⋅ 𝑏 𝐻 16 𝐹 , = ∫ τ 𝑑𝐴 𝐹 𝑠𝑙𝑖𝑑𝑒,𝑠 = μ ⋅ 𝐹 𝑁 Figure 4: EHL contact: Key Variables and Contact Forces velocity Parameter 𝑈 = 𝜂 0 ⋅ 𝑣 𝑠𝑢𝑚 𝐸 ⋅ 𝑅 (12) material Parameter 𝐺 = 𝛼 ⋅ 𝐸 (13) load parameter linecontact 𝑊 = 𝐹 N 𝑙 ⋅ 𝐸 ⋅ 𝑅 (14) load parameter pointcontact W = 𝐹 N 𝐸 ⋅ 𝑅 2 (15) isothermal filmheight parameter iso = cen,iso 𝑅 (16) the surface velocities v 1 and v 2 as well as the mean velocity v sum . Minimum and central lubricant film height h min and h cen are shown at their effective position. Both bodies are characterized by their radii R 1/ 2 and material parameters E 1,2 , ν 1/ 2 . The pressure peak acts in the area of the minimum lubricant film height h min . The central lubricant film height hardly changes over the entire contact width 2b and is therefore representative of the contact. This is required to calculate the shear stress, as this is described by the following relationship (see equation (10)) (10) The shear rate γ ˙ from equation (11) can be calculated with the assumption of linear velocity distribution and the calculated central lubricating film height by the relationship according to Lubenow [38]. (11) 2.3.2 Film thickness For the relative motion of a loaded contact, when a lubricant is present, a lubricant film builds up, which minimizes friction and wear. According to [39], the lubrication states can be classified as follows: • isoviscous-rigid, meaning rigid bodies and constant viscosity. • isoviscous-elastic, meaning deformable bodies and constant viscosity. • piezoviscous-rigid, meaning rigid bodies with viscosity depending on pressure. • piezoviscous-elastic, meaning deformable bodies with viscosity depending on pressure. In 1959, Dowson and Higginson introduced the dimensionless parameters for calculating the dimensionless isothermal film thickness H iso , which are listed in Table 1. The lubricant film thickness can be distinguished, as shown in Figure 4, into the central and the minimum film thickness. Their calculation using the dimensionless parameters differs only in the prefactors. The dimensiτ = η ⋅ γ̇ γ̇= 𝑣 1 − 𝑣 2 ℎ onless parameters are classified into the speed parameter U, the material Parameter G and the load Parameter W. Point/ Elliptical contact In ball bearings, raceway contacts predominantly occur as highly loaded, elliptical contacts in which the piezoviscous lubrication region dominates. Pure point contacts only occur under zero load conditions [40]. Equations for calculating lubricating film heights for pure point contacts are therefore not relevant for bearings. For elliptical contacts, Hamrock and Dowson provide equations in [41] for the central and minimum lubricating film height, assuming a flow direction along the shorter semi-axis. Chittenden and Dowson extend this approach in [42] by explicitly considering the flow direction and including different radii of curvature in and perpendicular to the flow direction (see equation (17)). This is particularly important when differential slip and bore slip overlap in the contact ellipse. This is the case, for example, in the raceway contacts of combined load deep groove ball bearings, angular contact ball bearings and the roller-rim contacts. (17) (18) On the other hand, Venner (point contact) [43] and Nijenbanning (elliptical contact) [39] take all lubrication areas into account. However, Nijenbanning makes the same assumption as Hamrock and Dowson regarding the direction of flow. Line contact The calculation of the minimum and central lubricant film thickness in line contacts is based on the same approximate equations as for point or elliptical contact. In 𝐻 , = 4.31 ⋅ 𝐺 0.49 ⋅ 𝑈 0.67 ⋅ 𝑊 −0.073 ⋅ (1 − x (−1.23 ⋅ (𝑅 𝑅 ) 2 3 )) 𝑅 𝑅 = 𝑅 𝑅 x ⋅ 𝑐𝑜𝑠 2 𝜉 + 𝑠𝑖𝑛 2 𝜉 𝑐𝑜𝑠 2 𝜉 + 𝑅 𝑅 x ⋅ 𝑠𝑖𝑛 2 𝜉 Science and Research 19 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 Table 1: Dimensionless parameter for film thickness calculation velocity Parameter material Parameter load parameter linecontact load parameter pointcontact isothermal filmheight parameter lows radial loads of up to 70 kN to be applied, misalignments or tilts to be adjusted, and axial loads to be introduced via a support bearing. The system is designed for rolling bearings with outside diameters of up to 180 mm and supports various lubrication concepts. For the present investigation, measurements with minimum quantity lubrication were used. In summary, the comparative data presented below are based on measurements taken using the test rig described in [47] for splash lubrication and the MEGT test rig [48] for minimum quantity lubrication. 4 Results To validate the model, comparing the friction torque resulting from the simulation of all internal bearing contacts with the experimentally determined values has proven to be particularly reliable. Since the friction torque depends heavily on the load distribution and the dynamic processes in the bearing, the total friction torque is used as the central comparison criterion for this purpose. To verify the simulation results of the FVA-Workbench, the friction torque is compared with experimental measurement data as well as with the results from Bearinx and the catalogue methods of Schaeffler and SKF. The data is colour-coded in the figures: the FVA-Workbench in blue, Bearinx in green, the Schaeffler catalogue in red, the SKF catalogue in cyan and the experimental measurements in black. Figure 5 shows a comparison of the friction torque for minimum quantity lubrication for the NJ216 cylindrical roller bearing in the speed range from 500 rpm to 5500 rpm. With minimum quantity lubrication, only the losses in the load zone and in the unloaded zone are taken into account. The bearing is subjected to purely radial loads with a load ratio of C/ P = 5. At 500 rpm, the result from the FVA-Workbench deviates from the measurement by approximately 20 %. As the speed increases, the friction torque in the simulation and experiment increases. In the range from 2000 to 4500 rpm, there is very good agreement. Above 4500 rpm, the deviations increase again. In the case of splash lubrication, Figure 6 shows the friction torques of the deep groove ball bearing 6319. The measurement data is taken from the final report [6] and refers to a purely radial load with a C/ P ratio of 6.5. The deviation of the FVA-Workbench is around 20 % at 500 rpm but is less than 5 % at 1500 rpm and 3000 rpm. Overall, the results of the FVA-Workbench correspond very well with Bearinx and the measurements. In both verifications, the Schaeffler catalogue method significantly overestimates the friction torque over the entire speed range. The result of the SKF catalogue method shows a large qualitative difference in the curve in Figure 5 but shows the best agreement with the measured values in Figure 6. The results of the FVA-Workbench in Figure 7 for the cylindrical roller bearing NU308 show the same tenden- Science and Research 20 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 highly loaded contacts as they occur in rolling bearings raceway contacts or on tooth flanks, the piezoviscous lubrication condition is predominant. While the Dowson and Toyoda equation [44] (see equation (19)) is only applicable to the piezoviscous region, all lubrication states are taken into account by the Moes [45] approach. (19) Correction Factors The calculation of the lubricant film height using the EHL theory is based on the assumption of isothermal conditions. The consideration of (shear heating) viscous heating in the inlet of the contact zone is neglected. However, the heating causes the viscosity and the lubricating film to decrease [46]. Murch and Wilson derive in [46] the equation for the thermal correction in the case of pure rolling in [46]. The correction factor is described as a function of the thermal load parameter (see equation (20)). The thermal load parameter L itself is given as a function of the dynamic viscosity at transmission pressure η 0 , the mean velocity v sum , the thermal conductivity of the oil λF and the temperature coefficient β. (20) (21) The thermally corrected central film height h cen can be calculated using equation (20) according to equation (22). (22) 3 Experimental determination of friction torque To verify the implementation, friction torques determined experimentally from two different test rig concepts are used. The first concept is based on a speed-controlled electric motor, a torque-measuring shaft and a test housing that incorporates both the test bearing and the lubrication system. Power loss is recorded via the measuring shaft, including losses from the support bearings as well as the test bearing. One drawback is the limited resolution, as the recorded friction losses cannot be exclusively assigned to the test bearing. Measurement data for splash lubrication is taken from the corresponding final report [47], which describes this test rig concept in detail. At the Institute of Machine Elements, Gears and Tribology (MEGT), a second test rig was developed that enables the friction torque of a single bearing to be measured directly [48]. The test bearing is mounted in a nearly frictionless hydrostatic bearing. The resulting circumferential force is measured via bending beams and converted into friction torque using the known lever arm. This setup al- = 3.06 ⋅ 𝐺 0.56 ⋅ 𝑈 0.69 ⋅ 𝑊 −0.1 Φ = 3.94 3.94 + 𝐿 0.62 𝐿 = η 0 ⋅ β ⋅ 𝑢 𝑚2 λ cen = Φ 𝑇 ⋅ ℎ 𝑐𝑒𝑛,𝑖𝑠𝑜 Science and Research 21 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 Figure 6: Comparison of measured and simulated friction torque results for splash lubrication Figure 5: Comparison of measured and simulated friction torques with minimum quantity lubrication Figure 7: Comparison of measured and simulated friction torque results for splash lubrication torque across different bearing types, speeds and lubrication conditions. It should be noted that the present investigations are based on individual comparisons. Nevertheless, even in this form, the results clearly demonstrate the superiority of the physically based universal contact friction model over catalogue-based estimations. 6 Conclusion and Outlook The universal contact friction model delivers highly accurate results in terms of simulated friction torque. This has been demonstrated by comparing the friction torque at minimum quantities and in splash lubrication. In contrast to empirical catalogue methods, the universal friction model enables a more detailed analysis of the physical contact conditions in rolling contact. In particular, the influence of geometric parameters on the friction torque can be specifically investigated. This allows optimization measures to be derived in bearing design, which lead to a reduction in friction and thus energy losses. The universal friction model thus makes an important contribution to the development of more efficient and durable rolling bearings and gear teeth. For comprehensive, holistic system optimization, the universal friction model is also applied to the gearing level. By extending the model to tooth contact, friction and loss mechanisms within the gear structure can be mapped more precisely. This enables improved evaluation of efficiency and optimized design of the overall system. The universal contact friction model provides the basis for: • determining the CO 2 e footprint during the service life of gear systems. • predicting the surface-induced damage. • calculating the electrical properties of individual contacts Science and Research 22 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 cies as those for the deep groove ball bearing 6319. At 500 rpm, the deviations between the FVA-Workbench and Bearinx results, and the measurements are approximately 40 %. However, at 1500 and 3000 rpm, the deviations are below 20 %. The SKF catalogue method particularly closely aligns with the measurements at 500 and 1500 rpm, and shows the same tendency as observed for the deep groove ball bearing 6319 at 3000 rpm. 5 Discussion The results demonstrate that the friction torque calculated using the universal contact friction model show a high degree of agreement with the experimental measurements across the entire speed range. Under minimum quantity lubrication, the deviation at 500 rpm is minimal, whereas under splash lubrication, differences mainly occur at low speeds. However, since the underlying measurements do not provide error bars, these deviations cannot be conclusively assessed. From medium speeds onwards, the simulation and experiment agree very well under both lubrication conditions. A simulation was performed for a deep groove ball bearing (6319) under a purely radial load with a C/ P ratio of 5, over the full speed range up to 4500 rpm under minimum quantity lubrication. This plot, which is based solely on the simulation (Figure 8), illustrates the model’s ability to capture all relevant lubrication regimes. The key distinction from catalogue methods lies in the modelling approach. While catalogue methods are based on simplified equations and cannot therefore account for different lubrication regimes, the universal friction model incorporates the results of EHL contact analysis. This enables high-resolution pressure and velocity distributions to be included in the calculation, providing a realistic representation of varying lubrication regimes. This results in a consistent, physically based description of friction Figure 8: Simulated friction torque: lubrication regimes • enabling, in the long term, the simulation of electrical equivalent models for gearboxes. • holistic system optimization of gearboxes 7 Acknowledgement This research was funded by Forschungsvereinigung Antriebstechnik (FVA) e.V., FVA project number 998 I. References 1. Sjovall, H. The Load Distribution within Ball and roller Bearings under Given External Radial and Axial Loads. Tek. Tidskr. Mek 1933, 9, 97-102. 2. Stribeck, R. Ball Bearing of Various Loads,Transactions of ASME, Vol. 29, 1907 pp 420-463. 3. Lundberg, G. Elastische Berührung zweier Halbräume. Forsch. Auf Dem Geb. Des Ingenieurwesens 1939, 10, 201-211. 4. Lundberg, G.; Palmgren, A. Dynamic Capacity of Roller Bearings. Acta Polytech. Mech. Eng. Ser. R. Swed. Acad. Eng. Sci. 1952, 2,96-127. 5. Palmgren, A. Ball and Roller Bearing Engineering; SKF Industries Inc.: Philadelphia, PA, USA, 1959. 6. Changenet, C.; Oviedo-Marlot, X.; Velex, P. Power Loss Predictions in Geared Transmissions Using Thermal Networks-Applications to a Six-Speed Manual Gearbox. Journal of Mechanical Design 2006, 128, 618-625, doi: 10.1115/ 1.2181601. 7. Holmberg, K.; Andersson, P.; Erdemir, A. Global energy consumption due to friction in passenger cars. Tribology International 2012, 47, 221-234, doi: 10.1016/ j.triboint.2011.11.022. 8. Holmberg, K.; Andersson, P.; Nylund, N.-O.; Mäkelä, K.; Erdemir, A. Global energy consumption due to friction in trucks and buses. Tribology International 2014, 78, 94- 114, doi: 10.1016/ j.triboint.2014.05.004. 9. Abdelbary, A.; Chang, L. Friction and wear. Principles of Engineering Tribology; Elsevier, 2023; pp 127-206, ISBN 9780323991155. 10. Stachowiak, G.W. How tribology has been helping us to advance and to survive. Friction 2017, 5, 233-247, doi: 10.1007/ s40544-017-0173-7. 11. Wikidal, F. Ritzelkorrektur: Programmbeschreibung (RIKOR F), Heft 481; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 1996. 12. Stiller, S.; Otto, M.; Stahl, K. Erweiterung Ritzelkorrekturprogramm (RIKOR) zur Bestimmung der Lastverteilung von Stirnradgetrieben: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 30 VII: Abschlussbericht, Heft 1077; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2013. 13. Weinberger, U. Erweiterung RIKOR: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 30 X: Abschlussbericht, Heft 1433; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2021. 14. Bong, H.B. Berechnung der Beanspruchungen und Sicherheiten von Stirnrädern mit der Methode finiter Elemente: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 127 IIa/ b: Abschlussbericht, Heft 322; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 1990. 15. Brecher, C.; Rieg, F. Realitätsnahe Berücksichtigung des elastischen Umfeldes auf den Zahneingriff mittels FEM: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 484 IV: Abschlussbericht; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2016. 16. Schroers, M.; Zahn, A.; Brimmers, J.; Glenk, C.; Brecher, C.; Rieg, F. Methode zurBerücksichtigung des interaktiven, quasistatischen Steifigkeitsverhaltens benachbarter Eingriffe und Doppelschrägverzahnungen in der Verzahnungsauslegung: Sachstandsbericht zum FVA Vorhaben: FVA-Nr. 377 III: Sachstandsbericht; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2020. 17. Lutz, M. Tragbildprogramm: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 525 I: Abschlussbericht, Heft 500; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 1999. 18. Sigmund, W. Schnecken-Schraubradgetriebe: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 452 II: Abschlussbericht, Heft 1163; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2015. 19. Roth, P.; Reißmann, J. Weiterentwicklung des Programmsystems SNESYS zur Berechnung von Schneckengetrieben: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 320 VIII: Abschlussbericht, Heft 1460; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2021. 20. Oster, P. EDV-Unterprogramm zur Berechnung der Steifigkeit und der Lebensdauer von Wälzlagern: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 364 I: Abschlussbericht, Heft 674; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2002. 21. Wang, D.; Jurkschat, T.; Otto, M. Low Friction, Lager2 (Wälzlager Reibungsberechnung) - Erweiterung der Berechnung der Wälzlagerreibung: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 701 I: Abschlussbericht, Heft 1157; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2015. 22. Zander, M.; Kehl, J.H. Erweiterung der Wälzlagerberechnung: Abschlussbericht zum FVA Vorhaben: FVA-Nr. 701 III: Abschlussbericht, Heft 1404; Forschungsvereinigung Antriebstechnik e.V: Frankfurt am Main, 2022. 23. Schaeffler. Bearinx Calculation Modules, 2025. 24. Epskamp, T.; Butz, B.; Doppelbauer, M. Design and analysis of a high-speed induction machine as electric vehicle traction drive. 2016 18 th European Conference on Power Electronics and Applications (EPE’16 ECCE Europe); Institute of Eletrical and Electronics Engineers, 2016; pp 1- 10. 25. FVA Software & Service GmbH. FVA-Workbench, 2025. 26. Höhn, I.B.-R.; Wirth, C.; Haefke, N. Design and optimization of automotive transmissions with the FVA-Workbench\textregistered. In Getriebe in Fahrzeugen 2011-Effizienzsteigerung im Antrieb Friedrichshafen, 07./ 08. Juni 2011, 2011. 27. Teutsch, R. Kontaktmodelle und Strategien zur Simulation von Wälzlagern und Wälzführungen. PhD thesis; Technische Universität Kaiserslautern, Kaiserslautern, 2005. 28. Schaeffler Technologies AG & Co. KG. Schaeffler Technologies AG & Co. KG, Ed., Wälzlagerpraxis: Handbuch zur Gestaltung und Berechnung von Wälzlagerungen (Antriebstechnik), 4th ed.; Mainz: Vereinigte Fachverl, 2015, 2015, ISBN ISBN: 978-3-7830-0401-4. 29. SKF: Rolling bearings, catalogue, PUB BU/ P1 17000/ 1 EN, 2018. Science and Research 23 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 41. Hamrock, B.J.; Dowson, D. Isothermal Elastohydrodynamic Lubrication of Point Contacts: Part III - Fully Flooded Results. Journal of Lubrication Technology 1977, 99, 264-275, doi: 10.1115/ 1.3453074. 42. Chittenden, R.J.; Dowson, D.; Dunn, J.F.; Taylor, C.M.; Johnson, K.L. A theoretical analysis of the isothermal elastohydrodynamic lubrication of concentrated contacts. I. Direction of lubricant entrainment coincident with the major axis of the Hertzian contact ellipse. Proceedings of the Royal Society of London. A. Mathematical and Physical Sciences 1985, 397, 245-269, doi: 10.1098/ rspa.1985.0014. 43. Venner, C.H.; Napel, W.E. Multilevel solution of the elastohydrodynamically lubricated circular contact problem Part 2: Smooth surface results. Wear 1992, 152, 369-381, doi: 10.1016/ 0043-1648(92)90133-s. 44. Dowson D. A Central Film Thickness Formula for Elastohydrodynamic Line Contacts, Elastohydrodynamics and Related Topics. Proc. 5th Leeds-Lyon Symp., 1978 1978, 60. 45. Moes, H. Optimum similarity analysis with applications to elastohydrodynamic lubrication. Wear 1992, 159, 57- 66, doi: 10.1016/ 0043-1648(92)90286-H. 46. Murch, L.E.; Wilson, W. A Termal Elastohydrodynamic Inlet Zone Analysis. ASME Journal of Lubrication Technology 1975, 212-216. 47. Thomas Jurkschat. Erweiterung von LAGER2 zur Dimensionierung von Wälzlagern in Industriegetrieben: Verlustleistung und Betriebstemperatur. Abschlussbericht FVA-Forschungsvorhaben 364 IV 2015. 48. Viktor, A. Detaillierte Beschreibung von Strategien und numerisch effektiver Methoden zur Dynamiksimulation vollrolliger Zylinderrollenlager. PhD thesis; Technische Universität Kaiserslautern, 2013. Science and Research 24 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0030 30. TIMKEN: Engineering manual bearings, catalogue, 2011. 31. ISO/ TR 14179-2: 2001(E): Gears - Part 2: Thermal loadcarrying capacity, 2001. 32. Wisniewski, M. Elastohydrodynamische Scmierung- Grundlagen und Anwendungen; Exper verlag: 71272 Renningen, 2000, ISBN 3-8169-1745-3. 33. Goksem, P.G.; Hargreaves, R.A. The Effect of Viscous Shear Heating on Both Film Thickness and Rolling Traction in an EHL Line Contact-Part I: Fully Flooded Conditions. Journal of Lubrication Technology 1978, 100, 346- 352, doi: 10.1115/ 1.3453183. 34. Krause, H., Poll, G. Mechanik der Festkörperreibung; VDI-Verlag GmbH, 1980. 35. Johnson, K. Contact Mechanics. Cambridge University Press, 1985. 36. Koch, O. Dreidimensionale Simulation von kombiniert belasteten Radialzylinderrollenlagern. PhD thesis; Ruhr- Universität Bochum, 2008. 37. Zhou, R.S.; Hoeprich, M.R. Torque of Tapered Roller Bearings. Journal of Tribology 1991, 113, 590-597, doi: 10.1115/ 1.2920664. 38. Lubenow, K. Axialtragfähigkeit und Bordreibung von Zylinderrollenlagern. Abschlussbericht FVA-Forschungsvorhaben 2002. 39. Nijenbanning, G.; Venner, C.H.; Moes, H. Film thickness in elastohydrodynamically lubricated elliptic contacts. Wear 1994, 176, 217-229, doi: 10.1016/ 0043-1648(94)90150-3. 40. Hamrock, B.J.; Dowson, D. Isothermal Elastohydrodynamic Lubrication of Point Contacts: Part 1 - Theoretical Formulation. Journal of Lubrication Technology 1976, 98, 223-228, doi: 10.1115/ 1.3452801. Motivation Lubricating oils should form a stable liquid film in rolling contacts and thus protect the surfaces of the contact partners and reduce friction and wear. Friction losses are generated, for example, in the rolling contacts of steel gears and rolling or sliding contacts of bearings. Those losses reduce the efficiency and increase warming. The heat, resulting from those friction losses, is dissipated into the environment and contributes to global warming. For example, current efficiencies of gearboxes in wind turbines, are said to be between 97.5 and 98 %. Even with the greatest efforts, the efficiency of large gearboxes is not exactly measurable. Also, the calculation options, related to gearing losses and bearing losses, are based on old approaches and use research results from that time [1, 2, 3]. The smallest improvements in efficiency of mass-produced products, such as wind turbine gearboxes, gearboxes in general, and motors will have a significant impact on overall global power losses and thus on global warming. As example: a 3-stage planetary gear unit with 2 planetary stages and one helical gear stage has more than 13 tooth contacts and more than 20 contacts in bearings, where frictional losses are generated. If the efficiency of this wind turbine gearbox could be improved by 0.2 % from 97.5 % to 97.7 %, the power gain at 5 MW/ h would be 10 kW/ h, and with a population of 40,000 gearboxes, this would result in 400 MW/ h. With 7,000 production hours per year, this represents an enormous value of 2,800 GW per year, which could be saved. Selecting lubricants with low friction coefficients for machinery will therefore make a significant contribution in reducing global warming. Goal setting The most important challenge of today is, reduce global warming. There are various possibilities to support this goal. A new approach to support the reduction of global warming is, reduce friction losses in all lubricated contacts. Therefore, it is necessary to agree specifications for lubricants with respect to their property of friction loss, by specifying the maximum allowable friction coefficient or moment and / or wear value. The goal is, to propose general limits for the friction behavior of lubricants in contact surfaces of gears and bearings. State of the art knowledge about friction losses in lubricated contacts Methods for calculating the heat generation and heat dissipation in gears were presented in [1] as early as Science and Research 25 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 Can lubricants with lower frictional torque in rolling contacts of bearings, gears and machines significantly reduce global warming? Dirk-Olaf Leimann* Lubricants for machines with machine elements as steel bearings and gears, such as used in gearboxes or engines, are supposed to reduce friction and wear in the rolling surface contacts, but, dependent on the lubricant, as shown below, they themselves generate lower or higher levels of friction, and thus energy losses, which influence global warming. This article compares various lubricants, as mineral and synthetic oils, to determine the differences of friction coefficients, wear values and frictional torque of those lubricants. It will be shown, how the appropriate selection of lubricants with lower friction or wear values can reduce power losses and thus have a positive impact on the climate and reduce global warming. Friction coefficients, wear values, and frictional torque are measured in tests on FE8 (bearing), FZG (gear) or two disk test rigs. The used data base in this paper contains 227 data sets with test results on gears (FZG, 6 oils), bearings (FE8, 41 oils) and two disk tests (11 oils), with overall more than 12000 data. Even, if the effect of reduced friction due to the lubricant selection may be small per individual bearing or gear, the total impact on global warming is enormous, given by the use in billions of machines and machine elements worldwide. Keywords energy loss, mineral oils, synthetic oil, friction coefficients, wear values, frictional torque Abstract * Dipl.-Ing. Dirk-Olaf Leimann Düsseldorfer Straße 4 47441 Moers ons regarding the properties and load behavior for the lubricant selection of rolling bearings and a few regarding the load behavior of gears. Examples of these can be found in [4, 5]. Tables 1, 2, and 3 show these accustomed requirements [4,5]. As visible from these specifications, no requirements for friction properties are present. Friction moment tests on FE8 and FZG test rigs FE8 tests for bearings and FZG tests for gears are very common to determine lubricant behavior. The images 1 and 2 show those test rigs. The purpose of the FE8 bearing tests and FZG gear tests, as shown in table 1, 2 and 3 is, to give results for lubricants with respect to wear and fatigue behavior. As already mentioned, no tests are recommended for friction behavior. Science and Research 26 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 1982. Methods for reducing losses in gear designs measures were described in [2, 3] in 1993 / 1994. All of these and other design measures have led to better efficiencies, being achieved in current gear designs. Due to the fact, that gear box designs, as mentioned before in the example wind turbine gear box, have more than 33 different lubricated contacts with different designs, sizes, load and speeds, the choice of the best lubricant for each contact leads at least to a high number of required lubricants with a veriety of viscosities and properties. But, in a gear box, only one lubricant can be used. Means, the properties may not be the best choice for all contacts. Therefore, a compromise in the choice of the final lubricant is mandatory. A lot of information for the selection of a lubricant for gears and bearings are given in [1, 11,16], where also tables and equations to calculate friction losses are given. However, to date, there are generally no specifications for the friction behavior of lubricants, i.e., friction coefficients or friction torques or wear values, when selecting lubricants for machines. There are many specificati- Property Test method Test conditions Criteria pass Oxidation ASTM D2893 95 °C < 6 % Oxidation ASTM D2893 120 °C < 6 % Foaming ASTM D892 3 x max. 75/ 10 Cupper corrosion ISO 2160 120 °C max. damage degree 2 Wear DIN 51819-3 DIN 51819-3 < 15 / 30 mg Ripplings DIN 51819-3 DIN 51819-3 no Micropittings DIN 51819-3 DIN 51819-3 no Rust, distiled water ISO 11007 ISO 11007 max. damage degree 1 Rust 0,5 % NaCl ISO 11007 ISO 11007 max. damage degree 3 Fatigue > 800 Wear < 30 mg Fatigue with water > 600 hrs Sludge slight Filter blokking no Wear of rollers < 15 mg Cage wear < 40 mg FE8 FAG Step 2 IEC 61400-4 FE8 FAG Step 4 Table 1: Criteria for selecting lubricants for rolling bearings, fresh oil [4] Standardized test methods for bearings acc. to IEC 61400-4 [5] bearing type Load Speed Temperature Runtime Roller wear Fatigue damage Rippling Micropitting P n t L kN rpm ° C h mg FE 8 stage 1 81212 100 7,5 80 80 < 30 small small FE 8 stage 2 81212 100 75 70 800 < 30 - - - FE 8 stage 4 with added water 81212 60 750 100 > 600 < 30 No - - Test method - - - - Table 2: Criteria for selecting lubricants for rolling bearings, fresh oil [5] Standardized test methods for gears acc. to IEC 61400-4 [5] Procedure name Test method Test conditions Recommended minimum requirement Gear wear (adhesive) FZG scuffing test ISO 14635-1 A/ 8.3/ 90 Fails > Load step 12 Gear wear (fatigue) FZG micropitting test FVA 54 / I - IV CGF/ 8.3/ 60 Fails > Load step 10 Gear wear (fatigue) FZG micropitting test FVA 54 / I - IV CGF/ 8.3/ 90 Fails > Load step 10 Table 3: Criteria for selecting lubricants for gears, fresh oil [5] Calculation methods for power losses and friction behavior In [1], equations, methods and examples are given for the calculation of power losses of gears, i.e. spur and helical gears, bevel gears, worm gears, rolling bearings, plain bearings, oil bath and seals and heat dissipation. There are calculation methods available to calculate friction coefficients and friction moments. These methods can be found in [1, 10, 11,16] The power loss P VZ of gears is the sum of load loss P VZP and no-load loss P VZ0 : (1) The load loss P VZP , can be calculated acc. to equation (2), where a friction coefficient µ mZ and a factor H V can be calculated according to [1, 10, 11]. H V is an important factor, which can also take the gear micro geometry into account. The calculation of the no-load loss P VZ0 contains a lot of influence factors and is difficult to determine. (2) The calculation method for bearings is like the calculation method of gears. The friction torque “loss” of a bearing T VL is the sum of the “load loss” T VLP and “no-load loss” T VL0 . (3) The load dependent friction torque T VLP is: VZ VZ = VZP ZP + VZ0 Z0 VZP mZ v VL VLP VL0 VZ VZP VZ0 VZP ZP = × mZ mZ × v VL VLP VL0 VZ VZP VZ0 VZP mZ v VL VL = VLP LP + VL0 L0 (4) with the load P 1 (Nm) and the bearing pitch diameter d m (mm). The no-load dependent friction torque T VL0 is: if (5) The no load dependent T VL0 friction loss is: if (6) For the detailed calculation equations and units, please refer to the equations, mentioned in [1, 10, 11,16]. For gears, the friction coefficient must be calculated, for bearings, friction coefficients are given in combination with a load dependent factor f 1 for different bearing types. See table 4. Friction coefficients vary between 0,0005 and 0,0040. Friction coefficient and f 1 values are dependent on source [1,10, 11,16]. To calculate the no-load friction moment for bearings, a f 0 factor is given for different bearing types and lubrication methods. See table 5. (f 0 values are dependent on source [1, 10,11,16]) These calculations and factors are later used in the tables for the comparison of the measured friction moments to the calculated friction moments. VLP LP = 1 × 1 × m VL0 0 (t) m (t) VL0 0 m (t) VLP 1 1 m VL0 L0 = 0 ×( (t) t) × ) , × m (t) VL0 0 m (t) VLP 1 1 m VL0 0 (t) m (t) × ≥ VL0 0 m (t) VLP 1 1 m VL0 0 (t) m (t) VL0 L0 = × 0 × m (t) VLP 1 1 m VL0 0 (t) m (t) VL0 0 m (t) × < Science and Research 27 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 Image 1: FE8 test rig [8,14] VZ VZP VZ0 VZP mZ v Image 2: FZG test rig [4] VZ VZP VZ0 VZP mZ v stress relationship, c, exponent in the stress-life equation, e, Weibull exponent, η a , hoop and residual stress factor, η b , lubrication factor, η c , contamination factor. The factor for the lubrication influence η b is calculated as: (8) with: ψ, bearing characteristic number, M, viscosity ratio related factor, κ, viscosity ratio, m, viscosity ratio related factor. The factors for equation 7 and 8 can be used from table 6 [12]. 1 0 a b c u b = × � , × � − �� Science and Research 28 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 New approaches for bearing life calculation In 1999, a new approach for the life time equation for bearings was published by E. Ioanides, G. Bergling and A. Gabelli [12]. The new equation included a new lubrication factor η b . The general life time equation is: (7) with : L 10 ,Basic rating life [Mrevs], A, scaling factor, C, Basic dynamic load rating [N], P, equivalent dynamic bearing load [N], P u , fatigue load limit ( see catalogue) [N], p, exponent in life equation, w, exponent in the load- 10 = � 〈 −� a × b × c × u � 〉 � × � � b Friction coefficients and determination of friction factor f 1 P / C = 0,1 - Min Max Deep groove ball bearing 0,0015 0,0030 0,0020 to be calculated 0,0009 (P 0 / C 0 )^0,55 Self-aligning ball bearing 0,0010 0,0030 0,0010 to be calculated 0,0003 (P 0 / C 0 )^0,4 Angular contact ball bearing single row 0,0015 0,0020 0,0015 (F/ C)^0,33 (40°) 0,0013 (P 0 / C 0 )^0,33 Angular contact ball bearing double row 0,0024 0,0030 0,0020 (F/ C)^0,33 0,0010 (P 0 / C 0 )^0,33 Cylindrical roller bearing 0,0010 0,0030 0,0005 1 0,0003 1 Needle roller bearing 0,0020 - 0,0005 1 - - Spherical Roller Bearing 0,0020 0,0030 0,0010 to be calculated 0,0035 1 Tapered roller bearing 0,0020 0,0050 0,0010 to be calculated 0,0005 1 Axial deep groove ball bearing 0,0012 - 0,0015 (F/ C)^0,33 0,0012 (P 0 / C 0 )^0,33 Axial spherical roller bearing 0,0030 - 0,0015 1 0,0006 1 Axial cylindrical roller bearing 0,0040 - 0,0035 1 0,0018 1 Bearing type Friction coefficient load dependent [1] f 1 Basic μ Factor load Friction coefficient general [exact determination see 10,16] μ Friction coefficient load dependent [exact determination see 10,16] f 1 Basic μ Factor load Table 4: friction values µ and load dependent friction factor f 1 Factor f 0 for no-load loss calculation for different bearing types and lubrication methods Min Max Min Max Min Max Min Max Min Max Deep groove ball bearing 1,5 2,0 1,5 2,0 3,0 4,0 0,7 1,0 0,7 1,0 Angular contact ball bearing single row 1,5 2,0 0,7 1,0 Angular contact ball bearing double row 3,0 4,0 1,6 2,0 Cylindrical roller bearing 2,0 3,0 2,0 3,0 4,0 6,0 1,0 1,5 1,5 2,0 Needle roller bearing 6,0 12,0 6,0 12,0 12,0 24,0 3,0 6,0 3,0 6,0 Spherical Roller Bearing 4,0 6,0 4,0 6,0 8,0 12,0 2,0 3,0 2,0 3,0 Tapered roller bearing 3,0 4,0 3,0 3,5 6,0 8,0 1,5 2,0 1,5 2,0 Axial deep groove ball bearing 1,5 2,0 1,5 2,0 3,0 4,0 0,7 1,0 0,7 1,0 Axial spherical roller bearing 3,0 4,0 3,0 4,0 6,0 8,0 Axial cylindrical roller bearing 2,0 3,0 - - 2,5 5,0 - f 0 2,0 4,0 1,0 4,0 8,0 2,0 Bearing type f 0 f 0 f 0 f 0 Oil bath / grease horizontal shafts [1] Oil bath / circulating [10] Oil bath / grease vertical shafts [1] Oil mist lubrication [1] Lubrication oil mist, drop, grease [10] Table 5: no-load dependent friction factor f 0 10 a b c u b Factor 0,1051 < κ < 0,41 use κ = 0,1051 if κ < 0,1051 0,41 < κ < 1 1 < κ < 4 use κ = 4 if κ > 4 Radial ball bearing Radial roler bearing Thrust ball bearing Thrust roller bearing M 0,87830 0,77860 0,77890 - - - m 0,05760 0,19090 0,07174 - - - ψ - - - 0,50 0,15 0,16 0,06 w - - - 1 / 3 1 / 2,5 1 / 3 1 / 2,5 p - - - 3 10 / 3 3 10 / 3 c / e - - - 9,3 9,2 9,3 9,2 A - - - 0,10 0,10 0,10 0,10 Table 6: Constants for the factors in the new approach for bearing life time calculation [12] Equation 8 was applied to the data for the axial cylinder roller bearing 81212 from table 8 for 6 different lubrication oils and different speeds, tested on a FE8 test rig. The diagram 1 shows the calculated factor η b and the measured friction moment with respect to the viscosity ratio κ. It is an interesting observation, that there is a relation between the friction moment and η b , unfortunately, there is no equation available to convert η b values to friction moments. The research continued in [17] with the “SKF Generalized Bearing Life Model”, where an equation for the surface risk function R S , as part of the general model, L 10GM , contains the influence of surface stress and conditions as lubrication, contamination, wear, and others. (9) S = ( , , c , 1 , 2 ) Data analysis friction behavior of lubricants from available research results In some research studies (references [6, 7, 8, 9]) with the goal to determine the load carrying capacity of gears and rolling bearings, friction coefficients, wear, and frictional torques values were measured and documented. In the cited papers, information about these lubricants is available on mechanical properties such as viscosity grade and, in some studies, also on chemical components, which results in total > 227 data cases. In [8], 4 different bearing types in combination with 10 lubrication oils and several load and speed combinations were examined and the frictional torque values were measured. The measured friction moment was compared with calculated friction moments acc. to the equations 4 to 6. The data and the results are shown in table 7. Table 7 shows, using data from [8] with the axial cylindrical roller bearing 81212 and gear information as an example, that with ten different lubricants, significant differences in the friction torques of the rolling bearing and the gear scuffing test results can be seen. From table 7 it can be assumed, that a direct influence on the friction moment by different chemical components and their amount is not visible. In [8], all lubricant tests were carried out with 4 bearing types: axial cylindrical roller bearing 81212, axial spherical roller bearing 29412, axial deep groove ball bearing 51212, and angular contact ball bearing 7312. A comparison of the Science and Research 29 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 S c 1 2 Friction moment Friction moment load calculated [1] Load Value Speed Temperature Hertzian Stress T T (Load Cal.) P C / P n t p H ν 40 ν 100 ϱ 15 S P Zn Ca Nm Nm kN rpm ° C N/ mm² mm²/ s mm²/ s kg/ dm³ ppm ppm ppm ppm 1 81212 [ 8 ] I1.M.A99.184 Mineral 1 5,30 0,87 12,2 5 30,0 80 1282 184 16,9 0,909 2680 996 6 57 12 2 81212 [ 8 ] I1.M.A99.184 Mineral 1 5,70 0,87 12,2 5 7,5 80 1282 184 16,9 0,909 2680 996 6 57 12 3 81212 [ 8 ] I2.P.SP.100 PAO 2 5,40 0,87 12,2 5 30,0 80 1282 100 14,0 0,860 4500 332 12 19 12 4 81212 [ 8 ] I2.P.SP.100 PAO 2 5,90 0,87 12,2 5 7,5 80 1282 100 14,0 0,860 4500 332 12 19 12 5 81212 [ 8 ] I3.M.SP.100 Mineral 3 8,60 0,87 12,2 5 30,0 80 1282 100 11,0 0,890 1900 115 0 24 12 6 81212 [ 8 ] I3.M.SP.100 Mineral 3 8,70 0,87 12,2 5 7,5 80 1282 100 11,0 0,890 1900 115 0 24 12 7 81212 [ 8 ] I4.M.PD.100 Mineral 4 8,60 0,87 12,2 5 30,0 80 1282 100 11,0 0,880 12200 1831 1178 546 12 8 81212 [ 8 ] I4.M.PD.100 Mineral 4 8,70 0,87 12,2 5 7,5 80 1282 100 11,0 0,880 12200 1831 1178 546 12 9 81212 [ 8 ] SG1.M.A20.146 Mineral 5 3,40 0,87 12,2 5 30,0 80 1282 146,5 14,5 0,895 26900 1322 16 32 12 10 81212 [ 8 ] SG1.M.A20.146 Mineral 5 5,10 0,87 12,2 5 7,5 80 1282 146,5 14,5 0,895 26900 1322 16 32 12 11 81212 [ 8 ] SG2.M.SP.79 Mineral 6 8,90 1,67 12,2 5 30,0 80 1282 79 9,8 k.A. 10400 489 0 275 no data 12 81212 [ 8 ] SG2.M.SP.79 Mineral 6 10,50 0,87 12,2 5 7,5 80 1282 79 9,8 k.A. 10400 489 0 275 no data 13 81212 [ 8 ] A1.M.SP.32 Mineral 7 8,50 0,87 12,2 5 30,0 80 1282 32 7,1 0,876 2600 54 0 819 10 14 81212 [ 8 ] A1.M.SP.32 Mineral 7 8,40 0,87 12,2 5 7,5 80 1282 32 7,1 0,876 2600 54 0 819 10 15 81212 [ 8 ] A2.M.SP.35 Mineral 8 9,40 0,87 12,2 5 30,0 80 1282 35 7,0 0,856 2300 188 0 52 8 16 81212 [ 8 ] A2.M.SP.35 Mineral 8 10,00 0,87 12,2 5 7,5 80 1282 35 7,0 0,856 2300 188 0 52 8 17 81212 [ 8 ] C2.M.SP.34 Mineral 9 7,70 0,87 12,2 5 30,0 80 1282 34 7,1 0,867 1100 466 697 678 11 18 81212 [ 8 ] C2.M.SP.34 Mineral 9 18,40 1,30 18,4 3,33 7,5 80 1282 34 7,1 0,867 1100 466 697 678 11 19 81212 [ 8 ] M2.M.ZP.106 Mineral 10 9,20 0,87 12,2 5 30,0 80 1282 106 11,8 0,896 10800 517 1166 4262 12 20 81212 [ 8 ] M2.M.ZP.106 Mineral 10 9,80 0,87 12,2 5 7,5 80 1282 106 11,8 0,896 10800 517 1166 4262 12 8,31 0,93 18,40 1,67 3,40 0,87 Maximum value friction moment Minimum value friction moment Count oils Oil mechanical data Oil chemical data Gear type: FZG- A/ 8.3/ 90 Load step fail > Average value friction moment Count cases Bearing type and size Source Test oil designation Oil type Table 7: Data for the frictional moment of 10 lubricants, measured on FE8 and FZG test and more detailed information about the lubricants 10 a b c u b 0,00000000E+00 5,00000000E-03 1,00000000E-02 1,50000000E-02 2,00000000E-02 2,50000000E-02 0,1051 0,1051 0,1051 0,1051 0,1051 0,1051 0,2841 0,2042 0,1732 0,1716 0,2360 0,1499 0,6078 0,4370 0,3706 0,3671 0,5049 0,3207 1,0805 0,7768 0,6587 0,6526 0,8976 0,5701 1,9208 1,3808 1,1710 1,1601 1,5956 1,0135 3,4146 2,4547 2,0817 2,0623 2,8364 1,8017 Lubrication factor η b [12] Friction moment x 1000 Nm [8] Viscosity ratio κ eta b Measured friction moment Nm x 1000 Diagram 1: Comparison of measured friction moments with calculated factor η b [12] Science and Research 30 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 Table 8: Measured friction moment and comparison with calculation results with data from [8] measured friction moment with the calculated friction moment acc. to the equations 4 to 6 for the bearings with the bearing load in table 7 and 8 clearly illustrates differences between calculation and measurement. The difference between measurement and calculation is big, near to 1000 %. In [9], six different lubricants with different viscosities and base oils were used. Table 9 shows results for gears and 2 different test procedures. Table 10 shows results from [9] for the bearings and gears, tested with the same or quite similar oils. Table 11 shows measured friction coefficients on a twodisk test rig with data from [7] and lubricant details. The data contain two chemical components and the viscosity data. Here it can be observed, that the measured friction values do not differ very much with respect to the speed. The question is, is this test suitable to gain friction informations. Could a FE8 bearing test also be represent for Gears? This question should be more examined, at least, both contacts have similar conditions regarding the frictional behavior. Diagram 2 gives some information to it. Probably the answer is no. Science and Research 31 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 Oil type short name Phosphorus Zinc ν 40 ν 100 ? 15 Test temperature Speed Hertian stress p c Water content fresh oil Test Wear Gear type mg/ kg mg/ kg mm²/ s mm²/ s kg/ dm³ °C 1/ min N / mm² ppm mg - Polyglycol PG1 2231 1689 224,1 39,5 1,059 60 2250 1723 1500 DGMK 575 7 / 5 C-GF Polyglycol PG2 2267 4 209,5 38,1 1,061 60 2250 1723 3000 DGMK 575 13 / 12 C-GF Polyalphaolifin PAO 470 2 224,5 27,9 0,859 60 2250 1723 120 DGMK 576 5 / 8 C-GF Ester E 136 5 199,9 24,8 0,950 60 2250 1723 300 DGMK 577 6 / 4 C-GF Mineral oil M 358 73 214,6 18,9 0,898 60 2250 1723 80 DGMK 578 9 / 9 C-GF Manual Transmission Fluid MTF 243 0 63,1 12,5 0,837 90 2250 1723 240 DGMK 575 16 / 13 C-GF Oil type Short name Phosphorus Zinc ν 40 ν 100 ? 15 Test temperatu re Speed Hertian stress p c Water content fresh oil Test Wear Gear type mg/ kg mg/ kg mm²/ s mm²/ s kg/ dm³ °C 1/ min N / mm² ppm mg - Polyglycol PG1 2231 1689 224,1 39,5 1,059 60 13 1853 1500 DGMK 377-1 38 C-PT Mineral oil M 358 73 214,6 18,9 0,898 60 13 1853 80 DGMK 377-1 58 C-PT Manual Transmission Fluid MTF 243 0 63,1 12,5 0,837 90 13 1853 240 DGMK 377-1 19 C-PT Results micro pitting test on FZG test bench according to DGMK 575 Results wear test on FZG test bench according to DGMK 377-1 FVA 488 Part Gears [ 9 ] Table 9: Detailed information to FZG tests [9] ν 40 VI ν 100 S P Zn Ca Mg mm²/ s mm²/ s ppm ppm ppm ppm ppm mg Nm [ 9 ] SKL PG 220 235 42 3254 2267 4 0 No Data FE 8 - 5,90 [ 9 ] SKL PG 220 235 42 3254 2267 4 0 No Data FE 8 - 5,90 [ 9 ] AZRL PG 220 235 42 3254 2267 4 0 No Data FE 8 - 9,75 [ 9 ] Gear PG1 224,01 231 39,48 No Data 2231 1689 0 0 DGMK 575 7 / 5 - [ 9 ] Gear PG1 224,01 231 39,48 No Data 2231 1689 0 0 DGMK 377-1 38 - [ 9 ] Gear PG2 213,84 236 39,9 No Data 2267 4 0 0 DGMK 575 13 / 12 - [ 9 ] SKL Ester 220 164 28 338 170 0 0 No Data FE 8 - 8,80 [ 9 ] Gear Ester 209,8 158 26,8 No Data 136 5 7 0 DGMK 575 6 / 4 - [ 9 ] SKL Mineral 220 101 19,5 10406 326 12 25 No Data FE 8 - 12,80 [ 9 ] AZRL Mineral 220 101 19,5 10406 326 12 25 No Data FE 8 - 16,00 [ 9 ] Gear Mineral 223,24 95 18,9 No Data 358 73 29 0 DGMK 575 9 / 9 - [ 9 ] Gear Mineral 223,24 95 18,9 No Data 358 73 29 0 DGMK 377-1 58 - [ 9 ] SKL PAO 220 163 28 2870 495 0 1 No Data FE 8 - 9,50 [ 9 ] Gear PAO 227,2 166 28,73 No Data 470 2 0 0 DGMK 575 5 / 8 - [ 9 ] SKL MTF 64 183 9,5 700 257 4 19 No Data FE 8 - 6,70 [ 9 ] Gear MTF 63,35 175 11,8 No Data 243 0 25 0 DGMK 575 16 / 13 - [ 9 ] Gear MTF 63,35 175 11,8 No Data 243 0 25 0 DGMK 377-1 19 - Source Oil data Test results Oil Type Oil mechanical data Oil chemical data Test name Wear Friction moment Bearing Type / Gear Table 10: Results from [9] on wear and friction moment for bearings and gears Nevertheless, there are ways to reduce friction coefficient values and friction torques and thus contribute to reducing losses and thus heat. Even the smallest improvements with the help of a suitable lubricant have an enormous effect on most of all applications. There might be also the opportunity to use coatings to additionally reduce the friction in rolling steel contacts. Due to lack of data, it was not possible to evaluate the influence of coatings. For both, bearings and gears, test methods for test benches such as FE-8 (bearings) or FZG (gears) are suitable to compare different lubricants about their friction behavior. It is recommended that these tests are included in the lubricant specifications for machines. For bearings, the recommendation for a lubricant specification would be to conduct an FE8 test with the conditions i.e. bearing type 81212, C/ P = 5, speed 30 rpm at 80° C and determine the frictional torque. The frictional torque for a lubricant should be less than 5 Nm. Diagram 3, which is the result of 168 measured friction moments, shows, that a limit of 5,4 Nm is feasible. For gears, a FZG test according to DGMK 575 is seen as suitable, in which the maximum wear for a lubricant should be less than 5-7 mg. Science and Research 32 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 Conclusion Measuring overall or individual efficiencies and power losses on machines, such as gearboxes and engines (combustion or electric motors), and drawing conclusions about friction coefficients or friction values for machine elements such as lubricants, gears or bearings is very difficult, almost impossible. The inaccuracies in the measurements are usually greater than the small individual differences in the values, to be measured. Table 11: Results for the frictional coefficient measured on a two-disk test Spur gear Path of contact Axial Bearings Roller width 0 D d m d > 0 0 > 0 >> 0 >> 0 A C E Friction Value μ Relative loss friction Moment Δ T with corrections Diagram 2: Frictional behavior of gears and axial cylindrical bearings Literature [1] D. Leimann, Wärmeentstehung und Wärmeabfuhr bei Getrieben, Firmenschrift PEKRUN, Iserlohn, 1982 [2] D. Leimann, Wärmearm konstruieren, Teil 1: Einfluss des Zahnflankenspiels auf die Erwärmung bzw. Verlustleistung von Zahnradgetrieben, antriebstechnik 32, 1993, Nr.3 [3] D. Leimann, Wärmearm konstruieren, Teil 4: Einfluss von Zahnbreite, Motorauswahl und Schmierstoff auf Erwärmung und Geräuschverhalten, antriebstechnik 33, 1994, Nr.4 [4] D. Leimann, Hansen selection criteria for lubrication oils for gearboxes in wind turbines Tagungsband17 th International Colloquium Tribology, TAE Esslingen, 2009 [5] NN, Wind Turbines - Part 4, Design requirements for wind turbine gear boxes, IEC 61400-4, Genève, 2012 [6] H. Surborg, Einfluss von Grundölen und Additiven auf die Bildung von WEC in Wälzlagern, Dissertation Universität Magdeburg, Shaker Verlag 2014 [7] Anatolij Smirnov et all, Wälzlagerermüdung bei Mischreibung in Abhängigkeit vom Schmierstoff, Vorhaben FVA 504 II, FVA Frankfurt, 2014 [8] T. Wolf et all, Einfluss des Schmierstoffes auf das Verschleißverhalten verschiedener Wälzlagerbauarten, Vorhaben FVA 327 II, FVA Frankfurt, 2007 [9] D. Brenner, J. Witzig et all, Zulässiger Wassergehalt in Getriebeschmierölen, insbesondere Polyglykol-Ölen und der Einfluss auf die Wälzlagerlebensdauer und die Zahnflankentragfähigkeit einsatzgehärteter Stirnräder, Vorhaben FVA 488, FVA Frankfurt, 2009 [10] Eschmann, Hasbergen, Weigand, Brändlein, Die Wälzlagerpraxis, zweite Auflage, Oldenburg Verlag München Wien, 1978 [11] NN, GfT Arbeitsblatt 5, Zahnradschmierung, Gesellschaft für Tribologie, Jülich [12] E. Ioanides et all, An analytical formulation for life of rolling bearings, Acta Polytechnica Scandinavica, The finnish Academy of Technology, Espoo, 1999 [13] NN, DIN/ ISO 281, Beiblatt 1, Wälzlager - Dynamische Tragzahlen und nominelle Lebensdauer - Lebensdauerbeiwert aDIN und Berechnung der erweiterten modifizierten Lebensdauer, DINMEDIA, Berlin, 2003 [14] N 060 mod 1, GfT Arbeitsgruppe „Datenbank Tribologische Prüfstände“ - Datenblatt FE 8 Prüfgerät, 2012 [15] NN, GfT Arbeitsblatt 7, Tribologie Definitionen, Begriffe, Prüfung, Gesellschaft für Tribologie, Jülich, 2002 [16] NN, GfT Arbeitsblatt 3, Wälzlagerschmierung, Gesellschaft für Tribologie, Jülich, 2006 [17] NN, SKF Group, PUB BU/ P9 15513 EN · March 2015 Science and Research 33 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0031 Diagram 3: Test results and limit value for frictional moment of FE8 bearing test 2 Materials and tribological phenomena 2.1 Materials This research looks at all materials used in automotive interiors. The spectrum includes leather and artificial leather, plastic components and complex elastomer profile seals in the door area. A comprehensive change in material development can currently be observed. In particular, the market for sustainable alternative materials to leather and artificial leather is seeing the continuous development of new variants of materials with properties that are as comparable as possible [1]. New material alternatives are also constantly being developed in the sealing sector in terms of elastomer composition and surface coatings. This involves a large number of new materials made from renewable raw materials. However, their physical and chemical properties differ from those of conventional materials, particularly on the surface. This has an influence on the resulting material properties, in particular on friction. 2.2 The stick-slip phenomenon Research on stick-slip phenomena spans both fundamental mechanics and applied tribology. Early reviews Science and Research 34 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0032 1 Introduction The prevention of background noise is becoming increasingly relevant in the modern automotive industry. It is not only a decisive factor for the perceived quality of the vehicle, but also plays a key role in creating a pleasant driving atmosphere. At a time when vehicles are more than just a means of transport, interiors are developing into complex comfort zones that appeal to all the occupants' senses. The traditional focus on visual and tactile aspects of interior design is now being expanded to include an acoustic dimension. In this context, disturbing noises caused by friction of materials are becoming particularly relevant. These noises, which are often subtle but certainly disturbing, can significantly impair the perception of vehicle quality, which is why they must be taken into account with due care during development and production. At the same time, the automotive industry is faced with the challenge of developing and using sustainable materials. The trend towards sustainability requires a change in the composition and properties of the materials used, which in turn can have an impact on their acoustic behaviour. In this area of conflict between comfort, quality and sustainability, automotive manufacturers (OEMs) and their suppliers are required to develop innovative solutions. The identification of material combinations that are free from disruptive noise sources and at the same time meet the requirements of sustainability and cost-effectiveness is a complex task. Particular attention must be paid to the stick-slip behaviour of materials, as this phenomenon is often the cause of unwanted noise. Overcoming these challenges requires not only a deep understanding of the underlying physical principles, but also the development and application of advanced testing and analysis methods. These methods must therefore be able to simulate and analyse the complex interaction of different materials under realistic conditions. Practical noise or stick-slip prevention in automotive interiors Martin Strangfeld, Susanne Fritz* The prevention of disturbing noises in the context of the automotive environment contributes to increasing vehicle quality and creates a pleasant driving atmosphere by minimising unwanted noise sources. The stick-slip behaviour of material combinations is of particular relevance here, as stick-slip is a frequent cause of disturbing noises and can be avoided through the targeted selection of suitable materials. There are already test methods for characterising the stick-slip behaviour, which are applicable in principle, but often have limitations in their significance. There is a clear trend towards increasingly realistic simulations of material combinations, taking into account real load scenarios. However, this poses a challenge for testing technology and the automated analysis of the results. Keywords Stick-slip, friction, real vehicle excitation, evaluation algorithms, friction force, automotive interior Abstract * Dr.-Ing. Martin Strangfeld Orcid-ID: https: / / orcid.org/ 0009-0003-3570-4286 Dr. rer. nat. Susanne Fritz Department Surfaces, FILK Freiberg Institute gGmbH Meißner Ring 1-5, 09599 Freiberg, Germany established the nonlinear dynamics of dry friction and stick-slip [2], with further analysis of mechanical stickslip vibrations through bifurcation and chaos theory [3]. More recent studies focus on polymers, linking deformation behavior to stick-slip [4], identifying distinct adhesive stick-slip modes [5], and characterizing tribological performance of EPDM under varying velocities [6]. The practical relevance is evident in engineering, where stick-slip underlies noise, vibration, and durability issues such as automotive buzz, squeak, and rattle [7]. Over the last few decades, noise in vehicle interiors caused by stick-slip behaviour has developed into a considerable economic problem, as noise is the cause of around ten percent of all complaints in the automotive sector. The stick-slip effect can be explained by comparing the frictional force and the force required to move the two bodies against each other from their rest position. This shows that the frictional force in motion is usually only slightly smaller than the displacement force mentioned above. If the sliding friction is significantly exceeded by the static friction, the energy required to overcome the static friction is no longer needed when sliding. As a consequence, the body is accelerated during sliding, which exceeds the acting force. As a result, the body reduces its speed and returns to its resting position until the acting force is sufficient to overcome the static friction. Figure 1 illustrates the changing behaviour of the frictional force F over time t during a stick-slip movement, taking into account the acceleration pulses a that occur when the static force is overcome. The characteristic acceleration pulses are analysed on a stick-slip test rig to identify the stick-slip behaviour. The dependence of the friction coefficient µ on the speed in a specific tribological system is shown in the form of a so-called Stribeck curve [9]. Figure 1 (right) shows the typical curve and the area with an increased risk of stick-slip. The phenomena mentioned are highly dependent on the material combination under consideration. 3 Current prevention of stick-slip behaviour Current research is evaluating various methods for assessing noise at different stages of vehicle development [10]. - Subjective assessment by test drivers is way of preventing disturbing noises. Here, a driver assesses the resulting noise in terms of its disruptive potential while driving the vehicle on a real road or a special test track with different road surfaces. However, this method can only be used at a late stage of development (prototype) or if problems occur. - Another option is noise measurement on the vehicle. Here, the jolting movement is stimulated by driving on the test track, using shaker systems or chatter rollers. The noises are recorded using objective methods, such as microphones, acoustic cameras or vibroacoustic measurements, and localised if necessary. - A different way of assessing noise is to measure the noise on the module. In an earlier development phase, it is possible to test individual modules in the same way as the entire vehicle, using special shaker systems in acoustic rooms. - The material pairing can be tested at a very early stage of development, whereby the material selection can be tested on the creaking or stick-slip test bench independently of the subsequent installation situation and without the need to produce prototypes. As part of this test, the noise is not measured directly, but the friction between the friction partners is assessed to determine whether stick-slip effects can potentially cause noise interference. Particularly in the initial phase of product development, material selection can be evaluated using a stick-slip test, which allows material combinations to be assessed in a controlled environment (temperature, relative hu- Science and Research 35 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0032 Figure 1: (left) Friction force and acceleration curve for stick-slip, (right) Stribeck curve [8] The stick-slip behaviour is quantified in accordance with VDA 230-206 using the so-called risk priority number (RPN) [11]. The calculation is based on the acceleration curve. As part of this approach, the height, area and frequency of acceleration peaks are evaluated. A grading scale was derived on the basis of the practical behaviour of the materials (see figure 3). Materials with a RPN between 1 and 3 are categorised as “acceptable” with regard to the stick-slip effect. Materials with a RPN between 4 and 5 are categorised as “conditionally acceptable”, while materials with a RPN between 6 and 10 are described as “unacceptable”. The frequency of pulses that occur over the travel distance is referred to as the pulse rate. This standard and the RPN scale enjoy a high level of acceptance worldwide and can be found in a large number of automotive group standards. For reasons of economy, VDA 230-206 was limited to the consideration of four different normal force/ speed combinations, although in reality a wide and varied load spectrum acts on the individual friction partners due to the stochastic jerking motion when driving. Especially Science and Research 36 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0032 midity), regardless of the final installation scenario. This approach saves resources that would otherwise be required for the production of prototypes. No direct quantification of noise is carried out as part of this process. The evaluation of the friction between the friction partners serves to identify potential noise disturbances that can occur as a result of stick-slip effects. Here, a specially designed friction test rig, the SSP 04 from Ziegler Instruments, is used to test the stick-slip effect. The test system is used to analyse the stick-slip tendency of material pairings, taking into account various influencing factors, including load, relative speed and climatic parameters. In this configuration, one test specimen is connected to a sled, while a second test specimen is attached to a steel spring (see Figure 2). The stick-slip test rig enables the precise determination of static and dynamic friction as well as stick-slip phenomena. The stick-slip tendency is determined on the basis of the interaction between the two test specimens. The data obtained with the test rig allows an evaluation of the stick-slip tendency of material pairings as well as an insight into their performance under different conditions. Figure 2: Measuring principle of the stick-slip test rig Figure 3: Background of RPN-calculation for the relative velocity range a real spectrum can have velocities up to 120 mm/ s. 4 Trends and current challenges When analysing the tribological behaviour of the material contacts in more detail, the excitation conditions primarily manifest the friction behaviour induced by the real vehicle movement. The friction behaviour based on stochastic movements is often not adequately represented by the application of the VDA conditions. This can lead to a misjudgement of the material behaviour and consequently to a large number of complaints. For this reason, a detailed analysis of the excitation has been carried out in recent years, which was then measured and evaluated under controlled conditions on the test bench. There are now group standards from OEMs that specify concrete stochastic signals for determining stick-slip. Up to now, however, the evaluation has been based on direct noise measurements. However, a material test in the laboratory always represents an abstraction of the real structure, so that the noise cannot always serve as an indicator. It is therefore necessary to extend the evaluation options to the resulting response friction signals at the same time as the process. Figure 3 provides an overview of the excitation conditions prevailing in the interior and their possible response signals. For the special case of elastomer profile geometries, the latest generation of the stick-slip tester has a wide range of equipment options available to take the special conditions of the seals into account. These include special specimen grips and the option of adjusting the angle of attack and overpressure of the seal in a targeted manner. However, gaskets are characterised by their high deformability and the resulting high damping constant. They also act like a kind of spring due to their elastic properties. These properties mean that if stick-slip effects occur Science and Research 37 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0032 Figure 4: Overview of the variety of different excitation and response signals (excitation is shown in blue in the figure, friction forces as response in orange) up the possibility of using various test benches and allows a more differentiated distinction between excitation and stick-slip signals. This allows the range of applications to be extended to a wider range of materials, load scenarios and measurement conditions. Two different algorithms were designed to quantify stick-slip: The primary algorithm is optimised for the analysis of high-frequency sampled friction force signals (sampling rate ≥ 1 kHz). After eliminating excitation-related oscillations, a theoretical acceleration signal is derived from the friction force signal via the relationship between the deflection path of the leaf spring and its change over time, from which a friction force-based RPN grade (FRPN) is then calculated following the original RPNcalculation. In standard-compliant tests, there is a high correlation between RPN and FRPN, with a correlation coefficient of 0.996 (see Figure 4 left). In 97 % of all measurements, the deviation is below the specified measurement uncertainty of the RPN. The innovative methodology now also allows the quantification of stickslip behaviour for damping materials, at increased excitation speeds, for variable movement profiles and on various friction test rigs. An alternative quantification approach was developed for friction test rigs with a lower sampling rate. Using easily extractable input variables, such as amount of force drops per millimetre and force drop height (see Figure 4 right), and applying the statistical method of multinomial logistic regression, approximate values for the FRPZ can be generated. A significant correlation to the RPN with a correlation coefficient of 0.973 can also be seen in standardised tests. The implementation of the FRPZ addresses a significant gap in preventive noise analysis and opens up perspectives for a more precise, more realistic and more versatile Science and Research 38 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0032 during the measurement, the acceleration pulses arriving at the acceleration sensor are significantly damped. Despite visible breaks in the friction force curve and/ or audible noises, only very low values are calculated for the risk priority number. In this respect, it should be mentioned that all other interior materials with a high damping effect have a similar property. The current generation of stick-slip test rigs is able to excite the sample with a stochastic motion profile. However, excitation-related acceleration peaks manifest themselves during such a measurement, which cannot be differentiated from the stick-slip-related acceleration pulses with the current evaluation. Consequently, when using the stochastic motion profile, it is currently not possible to calculate the RPN and therefore no valid stick-slip quantification is possible. 5 Current research 5.1 Friction force-based evaluation without the influence of acceleration In order to solve the problems of determining the RPN under stochastic excitation, a new method for determining the RPN based on the friction force has been developed (FRPN). When designing the algorithm, particular emphasis was placed on making the generated results compatible with the established RPN evaluation in the context of standard-compliant test procedures. In contrast to the conventional method, which is based on analysing acceleration curves, the newly developed algorithm focuses exclusively on evaluating friction force curves. Stick-slip phenomena are quantified on the basis of continuous force drops in the friction force values, whose frequency and amplitude are used as parameters. The preference for friction force curves opens Figure 5: Agreement of the FRPZ with conventional calculations on the left and an alternative approach for friction test benches with a low sampling rate on the right evaluation of friction instabilities. In addition, it is now possible to carry out the analysis for only a part of the friction measurement. This makes it possible to determine and analyse friction curves under continuous load variation in an experiment and thus determine stick-slip as a function of variation which has the potential to optimise the time and cost efficiency of complex test procedures. The algorithms developed were transferred into a software solution and are available for future stick-slip tests. 5.2 Friction Force drop identification algorithm - frictional distance to slip In vehicle development there are different approaches to prevent stick-slip. One of them is to reduce friction so that there are no occurrence of certain force drops of the friction force. The other one is to have a certain sticking behaviour and prevent the material combination from the transition from stick to slip [12]. In particular, for materials that exhibit pronounced deformability, such as elastomeric profile seals or highly elastic cushioning materials, the precise determination of the transition from the sticking to the sliding phase is of significant relevance. For this purpose, an automated detection algorithm was developed that functions independently of the specific excitation and quantifies the distance between the sticking and sliding transition (force drop) in soft materials. The algorithm developed is based on the identification of the exact point at which the gradient of the friction force curve either undergoes a significant change in sign or a substantial reduction (see Figure 6). The slope is determined with constant precision over a predefined data range. It is essential that the calculation is independent of the original excitation signal in order to eliminate potential influences from intrinsic noise and ensure precise analysis. The gradient or slope is analysed successively along the friction force signal. A change in the sign of the gradient indicates a transition between the tribological states of sticking and sliding within the friction pairing. Furthermore, a significant change is diagnosed if the gradient shows an abrupt reduction to at least 20 % of the initial value. In such cases, unstable friction behaviour is defined. The distance to the initial force drop is calculated as the difference between the position values at the point of gradient change and the last reversal point of the test rig’s axis of movement. The resulting distances can then be visualised in relation to the load parameters in order to analyse the influences of varying parameters and derive a worst-case scenario. The algorithm is characterised by its universal applicability to various forms of excitation in different test scenarios, whereby stochastic patterns are also taken into account in addition to linear and sinusoidal movement patterns. This ensures a high degree of realism in the analysis. The methodological innovation presented allows a detailed investigation of the influences of varying normal forces and speeds as well as an evaluation of climatic factors on tribological behaviour. The knowledge gained can be used to optimise the prediction models for the friction behaviour of complex vehicle systems, whereby acoustic phenomena such as squeaking noises can be minimised. Science and Research 39 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0032 Figure 6: Approach for the force drop identification algorithm References [1] Meyer, M., Dietrich, S., Schulz, H., Mondschein, A. (2021). Comparison of the Technical Performance of Leather, Artificial Leather, and Trendy Alternatives. Coatings, 11, 226. https: / / doi.org/ 10.3390/ coatings11020226 [2] Feeny, B., Guran, A. S., Hinrichs, N., & Popp, K. (1998). A historical review on dry friction and stick-slip phenomena. [3] Galvanetto, U., Bishop, S. R., & Briseghella, L. (1995). Mechanical stick-slip vibrations. International Journal of Bifurcation and Chaos, 5(03), 637-651. [4] Dong, C., Yuan, C., Bai, X., Qin, H., & Yan, X. (2017). Investigating relationship between deformation behaviours and stick-slip phenomena of polymer material. Wear, 376, 1333-1338. [5] Viswanathan, K., Sundaram, K., (2017). Distinct stickslip modes in adhesive polymer interfaces. Wear, (2017), 376. Jg., S. 1271-1278. [6] Sun, Q., Wang, S., & Lv, X. (2022). Research on stick-slip behavior and tribological properties of ethylene-propylene diene monomer under various wear velocity conditions. Polymer International, 71(8), 985-990. [7] Trapp, M., Chen, F. (2012). Automotive Buzz, Squeak and Rattle - Mechanisms, Analysis, Evaluation and Prevention, Elsevier [8] Klotzbach S. (2002). Ein nichtlineares Reibmodell für die numerische Simulation reibungsbehafteter mechatronischer Systeme; ASIM, Rostock. [9] Rorrer, R.(2000). A historical perspective and review of elastomeric stick-slip, Rubber chemistry and technology, 73(3): 486-503 [10] Moosmayr, T. A. (2009): Objektivierung von transienten Störgeräuschen im Fahrzeuginnenraum, Fortschritt-Berichte VDI, VDI Verlag, Düsseldorf [11] VDA - German Association of the Automotive Industry, VDA 230-206 (2021/ 10) - Examination of the stick-slip behavior of material pairs [12] Benhayoun, I., de Faverges, A., Bonin, F. et al. (2017). Less Interior Squeak and Rattle Noise Using a Simulation Driven Design Approach. ATZ Worldw 119, 36-41 Science and Research 40 Tribologie + Schmierungstechnik · volume 72 · issue 6/ 2025 DOI 10.24053/ TuS-2025-0032 6 Discussion and outlook The algorithms developed represent significant progress in the quantitative analysis of friction force curves for a wide range of interior materials. The versatile applicability with regard to various forms of excitation and their high precision open up new insights in the field of tribology and offer innovative approaches for the optimisation of vehicle components. The current development of test bench technology creates optimum conditions for the effective implementation of these algorithms. The current trend towards realising test conditions that are as close to reality as possible with the aim of achieving maximum informative value is leading to the design of test benches with synchronous excitation in all three spatial directions. This development requires an increased need for advanced analysis methods for the resulting complex data structures. The combination of realistic test conditions with precise data analysis opens up the possibility of a substantial improvement in preventive noise analysis. This methodological innovation has the potential to significantly reduce time-consuming and cost-intensive prototype measurements using shaker systems or comprehensive complete vehicle analyses. It allows more precise prediction and earlier prevention of acoustic phenomena, which can significantly increase both the efficiency of the development process and the quality of the end product. Future research approaches should focus on further refining the algorithms to handle even more complex data structures and on integrating these methods into holistic vehicle development strategies. Acknowledgement The presented results are part of the research projects (“Dist2Slip - determination of the distance to first slip”, 49MF210118 as well as “Stick-slip-quantification through friction force”, 49MF180089) and partly funded by the German Federal Ministry for Economic Affairs and Energy (BMWi) based on a resolution of the German Bundestag via the project management organization EuroNorm GmbH. Gratitude is expressed for the provided support. 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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! 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