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Tribologie und Schmierungstechnik EDITOR IN CHIEF MANFRED JUNGK 1 _ 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 1 | 2025 Volume 72 Editor in chief: Dr. Manfred Jungk Tel.: +49 (0)6722 500836 eMail: jungk@verlag.expert www.mj-tribology.com Editorial director: Ulrich Sandten-Ma Tel.: +49 (0)7071 97 556 56 / eMail: sandten@verlag.expert Editor: Patrick Sorg Tel.: +49 (0)7071 97 556 57 / eMail: sorg@verlag.expert Dr. rer. nat. Erich Santner Tel.: +49 (0)2289 616136 / eMail: esantner@arcor.de Contributions marked with the author’s initials or full name represent the author’s opinion, not necessarily that of the editorial office. We take no responsibility for unsolicited contributions. The author is responsible for obtaining the rights to pictures. 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ISSN 0724-3472 ISBN 978-3-381-13771-8 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 1/ 2025 DOI 10.24053/ TuS-2025-0001 At the European Council meeting of 12 December 2019 The Netherlands, Danish, Luxembourg and Swedish delegations presented information on risks related to polyand perfluorinated alkylated substances (PFAS). What happened since then? The industry faces a EU Restriction proposal of PFAS under the EU REACH Regulation and started to voluntarily phasing out fluorosurfactants, which are the PFAS substances under the most intense spotlight. As the current restriction proposal on PFAS is much wider and may restrict some applications that are critical to society with a huge impact on many downstream products in use in our daily lives. PFAS are a large class of thousands of synthetic chemicals and are increasingly detected as environmental pollutants and some are linked to negative effects on human health. As I learned early in my career when visiting a lead engineer at a car OEM from his printed statement behind his desk, “in god we trust, all others bring data”. Hence, I found two related information portals. In the Collaborative Research Center (CRC) 1349 “Fluorine-Specific Interactions”, 50 PhD students and Post- Docs are conducting research in the field of fluorine chemistry in 20 working groups of the Freien Universität Berlin, the Humboldt University Berlin, the Technical University Berlin, the Federal Institute for Materials Research and Testing and the Leibniz Research Institute for Molecular Pharmacology on fluorochemical issues. The scientific objectives are to understand and control the complex interactions that can emanate from fluorinated structural units in chemical systems. This will shed light behind the European Chemicals Agencies (ECHA) position on “negative effects on human health”. On the topic of “increasingly detected as environmental pollutants” the OECD publishes series or reports on Risk Management of Chemicals. One report presents a synthesis of publicly available information on perfluoropolyethers (PFPEs), with the aim of elucidating the identities of PFPEs on the global market and analyzing their life cycle. This includes their production and use, presence of other PFASs as impurities in commercial formulations, degradation mechanisms, and environmental releases of PFPEs and other PFASs present in commercial formulations. Even though in discussions I had over the last couple of years that e.g. for bearings PFPE might be easier substituted than the sealing material PTFE I think that report is most relevant for the lubricant society in general. Indeed, during ECHA’s recent meetings of their Committees for Risk Assessment (RAC) and for Socio-Economic Analysis (SEAC) they progressed their evaluation of the proposed restriction on PFAS. The committees reached provisional conclusions for applications of fluorinated gases and announced the sectors they will evaluate in their upcoming meetings in June include Medical devices, Lubricants, Transport, Energy and Electronics and semiconductors. In my earlier career I had dealt with PTFE in Anti-Friction Coatings and proposed PFPE fluids as wind turbine gear box oil, so remember Tribology is everywhere. Your editor in chief Manfred Jungk PFAS? Events 2 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 Events We look forward to your contribution! The scientific journal Tribologie und Schmierungstechnik (TuS) is one of the leading publications for tribological research in Germany, Austria and Switzerland. As the official journal of the Society for Tribology (GfT) in Germany, the Austrian Tribological Society (ÖTG) and Swiss Tribology, the issues provide information on research from industry and science, current events and developments in the specialist community. Further information on the journal and publication: https: / / elibrary.narr.digital/ xibrary/ start.xav? zeitschriftid=tus&lang=en Date Place Event ► 26.04. - 29.04.25 Copenhagen, Denmark ELGI 35 th Annual General Meeting ► 13.05. - 15.05.25 Brannenburg, Germany Oildoc Conference ► 18.05. - 22.05.25 Atlanta, Georgia (USA) 79 th STLE Annual Meeting & Exhibition ► 28.07. - 30.07.25 Zürich, Switzerland European Conference on Tribology - ECOTRIB ► 29.09. - 01.10.25 Wernigerode, Germany 66. GfT Conference Tribology Contents 3 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 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 1 April 2025 5 Felix Schlegel, Marius Hofmeister, Dino Osmanovic, Katharina Schmitz Swelling, wear and property changes of high-performance polymers in oil-hydraulic tribological contacts 14 Jan Euler, Georg Jacobs, Timm Jakobs, Thomas Decker, Noah Smeets, Julian Röder Influence of wear and manufacturing inaccuracies on the performance of a conical plain bearing main bearings for wind turbines 24 Dieter Mevissen, Sebastian Sklenak, Christian Brecher, Thomas Bergs Elastic-plastic micro contact calculation for large scale, lubricated rolling-sliding contacts 33 Carsten Heine Oxidation Index - Determining the Ageing of Industrial Lubricants 1 Editorial PFAS? 2 Events Science and Research 42 News Gesellschaft für Tribologie Swiss Tribology Columns Preface For authors Authors of scientific contributions are requested to submit their manuscripts directly to the editor, Dr. Jungk (see inside back cover for formatting guidelines). Anzeige 4 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 BOOK RECOMMENDATION expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 \ 72070 Tübingen \ Germany \ Tel. +49 (0)7071 97 97 0 \ info@narr.de \ www.narr.de This monograph takes a new look at tribology with its basic concepts of friction and wear using the example of lubricating greases. The consideration of the phenomenon of occurring instabilities and the introduction of the entropy concept into lubricating grease tribology provide a new perspective on known phenomena. The second part of this book presents a wide range of experimental possibilities for investigating lubricating greases. Contents Introduction to Instability and Postmodern Tribology - On the Phenomenon of Self - Organization - Postmodern Grease Tribology - Lubricating Grease - Rheological behavior of Lubricating greases - A Selected Traditional Wear Model - The Extension of the Wear Concept Erik Kuhn On the Tribology of Lubricating Greases An energetic approach to post-modern tribology Tribologie - Schmierung, Reibung, Verschleiß 1st edition 2025, approx. 210 p. €[D] 118,00 ISBN 978-3-381-14171-5 eISBN 978-3-381-14172-2 1 Introduction and state of the art This article examines whether tribologically optimized high-performance polymers can replace non-ferrous metal alloys in oil-hydraulic tribological contacts. In the first chapter these contacts are classified and delimited, followed by a presentation of currently used materials and the current state of research on polymers in hydraulic contacts. Building on this, methods for characterizing selected polymers are presented, followed by associated test results, that are discussed in the context of the research question afterwards. 1.1 Tribological systems in oil-hydraulic machines Hydraulic drives are used when large forces and torques or high power in a limited space are needed, which mechanical or electro-mechanical solutions cannot achieve [1]. Hydraulic drives are suitable for that, because power is transmitted by a fluid, allowing force and torque transmission via surface and volume ratios without solid surface contact [1]. Therefore, the highest tribological loads in hydraulic systems occur at the input and output, where hydraulic and mechanical energy are converted into each other. Rotary machines typically used for this experience higher tribological loads than linear machines. Figure 1 classifies the tribological loads of various rotary hydraulic machine types based on data from Wegner [2], Gärtner [3], and Pietrzyk [4], supplemented in an overview by Czichos [5]. As some rotary machines operate under variable load and speed, only the upper limit of typical loads is clearly defined in Figure 1. Hydraulic vane machines are less frequently used, and gear machines are tribologically similar to gearboxes. Thus, afterwards the focus is on piston pumps, whose tribological contacts are unique for hydraulics. A basic distinction is made between radial, axial and in-line piston machines [1]. As their tribological contacts do not differ fundamentally, the axial piston machine (Figure 2), the most common in industry and research, is used here as case study. The operating principle of an axial piston machine is explained in detail by Ivantysinova [6]. The machine functions as both a pump and a motor. In pump mode, oil is drawn in and expelled by an oscillating piston movement, generated by the rotation of the cylinder block and the pistons’ support on the swash plate via slippers. The three key tribological contacts are the piston-bushing, cylinder block-valve plate, and slipper-swashplate contact [6]. They share the following characteristics: Science and Research 5 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 Swelling, wear and property changes of high-performance polymers in oil-hydraulic tribological contacts Felix Schlegel, Marius Hofmeister, Dino Osmanovic, Katharina Schmitz * submitted: 23.08.2024 accepted: 6.02.2025 (peer review) Presented at GfT Conference 2024 In oil-hydraulic tribological contacts, hard-soft material pairings of steel and non-ferrous metal are usually used, whose main component copper is often alloyed with lead. These materials have significant disadvantages in terms of oil ageing and toxicity, which is why research is being carried out into alternatives. This article examines whether modern high-performance polymers can replace conventional non-ferrous metal alloys in oil-hydraulic contacts. Experimental studies on the media compatibility of various PEEK, PAI, PI and POM-H polymers in mineral and ester oil are discussed. In addition, the friction and wear behavior of PEEK and PAI is quantified by tribometer tests under typical oil-hydraulic conditions. Keywords hydraulics, alternative materials, media compatibility, wear, polymer Abstract * Felix Schlegel, M.Sc. (corresponding author) Orcid-ID: https: / / orcid.org/ 0009-0004-5501-3754 Marius Hofmeister, M.Sc. Orcid-ID: https: / / orcid.org/ 0000-0002-6672-5800 Katharina Schmitz Univ.-Prof. Dr.-Ing. Orcid-ID: https: / / orcid.org/ 0000-0002-1454-8267 Institut für fluidtechnische Antriebe und Systeme der RWTH Aachen University Campus-Boulevard 30, 52074 Aachen Dino Osmanovic B.Sc. For the following discussion, the slipper-swashplate contact was chosen as a representative example. The comparatively complex slipper kinematics are discussed in detail by Manring [8] and are briefly outlined here. The slippers are connected to the pistons via ball joints, providing three degrees of rotational freedom, allowing them to follow the swash plate under the piston load. Due to centrifugal forces, friction and hydrodynamics, the slippers tilt along the path tangent and radius and rotate around their axis, creating an uneven gap with an average height of 2 - 20 µm [7, 8]. The main slipper load is the piston pressure force, oscillating with the change of suction and delivery pressure at twice the pump’s rotation frequency. Additionally, inertia, friction and centrifugal forces occur [8]. To reduce the slippers nominal load, usually several kilonewton, the slippers are hydrostatically relieved by feeding oil from the piston chambers into the lubrication gaps through bores in the pistons and slippers (Figure 2 left). It is possible to adjust the contact’s remaining normal force by the area ratio of the piston cross-section to the compensated slip- Science and Research 6 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 • High normal forces due to support of hydraulic pressure (> 50 kN in large pumps) • High relative speeds compared to plain bearings (up to 20 m ⁄ s) • Comparatively large conformal surface contacts • Elasto-hydrodynamic lubricated (EHL) contacts with high deformation • Lubrication with hydraulic fluid under full immersion with comparatively large quantity of fluid, that is permanently filtered and cooled • Typical oil temperatures of approx. 40 - 80 °C • Pronounced load and speed-dependent contact pressure distribution due to deformation and complex kinematics [7] • Hydrostatic relief and hydrodynamic pressure build up relative velocity [m ] 1 10 100 1000 0.1 0.01 0.001 0.01 0.1 1 10 100 1000 10000 hertzian surface pressure [N/ mm 2 ] turbine ring seals valve guides slide ring seals aerodynamic spring bearings joint bearings gearboxes ball bearings sliding bearings brakes/ clutches piston rings friction gearboxes piston pumps vane pumps gear pumps Figure 1: Classification of oil-hydraulic tribological systems (based on [5]) suction line delivery line slipper piston cylinder block valve plate drive shaft swashplate bushing hydrostatic relief volume flow compensated slipper area sealing land Figure 2: Structure of an axial piston pump per area. With usual relief degrees of 80 -105 % [6, 8], contact pressures of approx. 0.01 - 3 MPa typically result at relative speeds of up to 20 m/ s. Slippers are usually designed to maintain a remaining normal force to close the lubrication gap and prevent floating, which can cause severe edge wear and high leakage losses. This load adjustment, along with lubrication and cooling from hydrostatic relief, provides the basis for exploring polymers in this unusual application. A key challenge will be the normal force adjustment to polymers while preventing floating. 1.2 Materials for oil-hydraulic tribological systems In oil-hydraulic tribological contacts, self-adjusting hard-soft pairings of non-ferrous metal and hardened steel are commonly used. These allow the softer contact partner to run-in, embed particles, and reduce adhesion risk. Typical used non-ferrous metals are special brass alloys (e.g. CuZn37Mn3Al2PbSi), gunmetal (e.g. CuSn7ZnPb), and bronze alloys (e.g. CuPb15Sn). Special brass alloys, containing 20 -30 m% zinc, tin, silicon, about 10 m% additional elements like aluminum, manganese, nickel, and iron, as well as 0.1 - 2 m% lead, are widely used [9]. A popular example is CuZn40Al2Mn2Si, known as Aeterna 3838. Lead improves grain refinement, machinability, corrosion resistance, and emergency running but is toxic to humans and the environment, classified as such under CLP and REACH regulations. Lead-free copper alloys like CuZn30Al2Mn22Ni1FeSiSn (OF2299) are available but not widely accepted by component manufacturers [10]. Furthermore, copper catalyzes oil aging and increasing acid content with oil age can cause copper corrosion and component damage [11]. That’s why, polymers may be an alternative to avoid these issues. In a “disc-on-disc” experiment with HLP46, on the tribometer described in chapter 2.2, at 40 °C of fluid temperature, 3 MPa nominal pressure, 15 m/ s relative speed and 216 km wear distance, the sliding friction- (f dyn ) and wear coefficients (K W ) for special brass alloys were measured by the authors (Table 1). 1.3 State of research of polymers in oil-hydraulic machines Polymers are typically classified based on mechanical properties and maximum operating temperature as standard, technical and high-performance polymers [12]. Previous Research on polymers in water-hydraulic piston pumps, e.g. by Rokala [13], Schoemacker [14], Li et al. [15, 16], Guan et al. [17] and Ni et al. [18], indicates, that only high-performance polymers, particularly PEEK, are suitable for hydraulic tribological contacts. Therefore, the authors selected PEEK natural, PEEK-CF30 (30 vol% carbon fibers), PEEK HPV (10 vol% carbon fibers, 10 vol% graphite, 10 vol% PTFE), PAI natural, PI natural and POM-H to investigate their suitability in oil-hydraulic contacts. These polymers have different structures. POM-H and PEEK are semi-crystalline thermoplastics, while PI and PAI are amorphous thermoplastics, also known as “pseudo thermoplastics”, due to their decomposition temperature being lower than their re-melting temperature. Amorphous thermoplastics have random arranged polymer chains, whereas semi-crystalline thermoplastics form regularly crystalline areas. The spaces between these chain structures allow fluid molecules, additives and fillers to be incorporated. Thus, thermoplastics are generally swellable and soluble [12]. Extensive databases for these polymers are available from manufacturers and research, which are not reviewed here. The density ρ, friction and wear properties of the selected polymers, based on ISO 7148-2-A-dr-F4-U4 dry “pin-on-disk” tribometer experiments are shown in Table 2 [19-24]. To the best of the authors’ knowledge, there is no data for characterizing the friction and wear behavior of these polymers under typical oil-hydraulic operating conditions, measured in experiments with comparable geometries for piston pumps. Moreover, the impact of hydraulic fluids on the properties of these polymers is not well understood. This article aims to address these gaps by presenting insertion and tribometer tests of the selected polymers to evaluate their suitability for oil-hydraulic tribological contacts. Science and Research 7 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 Material (0.5 m/ s / 12.5 m/ s ) [ / ] [ mm 3 N m ⁄ ] CuZn40Al2Mn2Si 0.14 / 0.032 0.95 ∙ 10 − 10 CuZn30Al2Mn2Ni1FeSiSn 0.14 / 0.05 0.98 ∙ 10 − 10 Table 1: Sliding frictionand wear coefficient of special brass alloys Material [ / ] [ mm 3 N m ⁄ ] [g/ cm 3 ] -limit [ MPa ∙ m s ⁄ ] PEEK [19] 0.3-0.5 9,33 ∙ 10 − 6 1.31 0.33 PEEK-CF30 [20] 0.2 − 0.3 6.66 ∙ 10 − 7 1.40 - PEEK-HPV [21] 0.15 − 0.25 6.66 ∙ 10 − 7 1.45 0.66 PAI [22] 0.35 − 0.6 1.66 ∙ 10 − 6 1.41 0.32 PI [23] 0.25 − 0.5 1.0 ∙ 10 − 6 1.50 2.9 POM-H [24] 0.2 − 0.3 2.66 ∙ 10 − 6 1.50 0.26 Table 2: Dry sliding frictionand wear coefficient of different polymers on steel ing on the density ρ r of the sample liquid and the gravitational constant g (eqn. 1). eqn. 1 The pressure ∆p acts on the bottom surface area A 0 and leads to an additional weight force F g which is proportional to the measured mass m V (eqn. 2). eqn. 2 The volume of the sample material V s equals the change of liquid level multiplied by the difference of A 0 and the cross-section area A 1 of the support (eqn. 3). eqn. 3 2.2 Tribometer tests - friction and wear To investigate friction and wear, a specialized “Slipperon-Disc” Tribometer, shown in Figure 4 was developed, to observe a simplified slipper-swashplate contact at nominal contact pressures of up to 30 MPa and relative velocities of up to 15 m/ s. Five abstracted slippers, each with only one ring surface, with an outer diameter of 16 ± 0.05 mm and an inner diameter of 8 ± 0.05 mm are mounted in a slipper holder at a mean frictional radius of 40 mm. To mitigate tolerances, the slippers were surface grinded and lapped together when installed. There is no hydrostatic relief. Instead, the test is conducted fully immersed in approx. 0.7 l of HLP46 at a controlled fluid temperature of 40 °C. = r g g g = 0 = r 0 g s = ( 0 1 ) = 0 1 0 r Science and Research 8 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 2 Methodology and experimental setup The accuracy and the standard deviation of the used measurement instruments in their actual lineup, is given by Table 3. The instruments for the insertion test and the optical measurements were used at 23 °C ± 2 °C. A separate precision scale in a clean room classified as DIN EN ISO 14644-1 class 7/ 8 was used to determine wear masses. Optical inspections of test objects were conducted at x50 magnification, while roughness measurements were performed using a focus variation microscope at 100x magnification. 2.1 Insertion tests - media compatibility To investigate media compatibility, two fluids were chosen following DIN 51524-2: a mineral oil-based hydraulic fluid HLP46 (ρ: 0.878 g/ cm3) and a synthetic ester oil HEES46 (ρ: 0.922 g/ cm 3 ). According to DIN EN ISO 175: 2011-3 and DIN ISO 1817: 2016-11, 3 samples of each of the 6 polymers were fully immersed in 75 ml of fluid at 40 °C ± 2 °C. The samples sizes were Ø 18 x 5 mm, differing from the standard. After cleaning with isopropanol, hardness (ISO 48-4), mass, volume and optical appearance were assessed over 42 days at intervals specified in DIN EN ISO 175: 2011-3. Figure 3 illustrates the setup used for volume determination. The measurement setup comprises a vessel on a precision scale and a support anchored on the ground on one side, extending into the vessel on the other. Initially, the vessel is filled with a reference liquid of known density, and the scale is zeroed. Next, the sample material is placed on the support, causing the liquid level to rise (∆h) and thereby increasing the hydrostatic pressure ∆p, depend- Measured Value Measurements per Sample Accuracy/ Accuracy Class Standard Deviation 4 0.02 mg 0.03 mg 3 0.01 mg 0.016 mg H 5 0.24 Micro-Shore D continuously 0.2 % (HBM) continuously 0.03 % (Bosche) n 1024 ppr. ± 0.5 % - Roughness 5 0.01 μm 0.002 μm Measured Value Measurements per Sample Accuracy/ Accuracy Class Standard Deviation 4 0.02 mg 0.03 mg 3 0.01 mg 0.016 mg H 5 / 0.24 Micro-Shore D continuously 0.2 % (HBM) continuously 0.03 % (Bosche) n 1024 ppr. ± 0.5 % - Roughness 5 0.01 μm 0.002 μm Table 3: Key data of the used measurement instruments 1 2 g g A 1 A 0 vessel support ∆h Figure 3: Measurement setup for the determination of sample volume Load application is facilitated via an air bearing, connected to a force sensor. Relative velocity is determined by the rotational speed of then drive shaft, while friction torque is measured by a force sensor through the support of the rotor holder. The measurement of wear masses is sensitive to the slipper cleaning process, hence an ultrasonic cleaner with isopropanol is used. It was observed that the cleaning process is robust, but particles stored in the polymer due to contaminated test fluid can significantly affect measurements. The particle load in the test fluid was classified as 18/ 15/ 11 according to ISO 4406: 1999. Furthermore, the running surfaces of the slippers were optically examined at 50x magnification to exclude larger particles before weighing. Wear coefficients are calculated using eqn.4 incorporating the measured test time t, the nominal contact pressure p C , the relative velocity v r , the polymer density ρ, the contact surface area A C , and the wear mass m W according to Czichos [5]. eqn. 4 3 Experimental results This section first shows the experimental results of the insertion tests and then the results of the friction and wear tests, followed by a discussion and comparison in section 4. 3.1 Insertion test results The results of the insertion tests for HLP46 are shown in Figure 5 and the results for HEES46 in Figure 6. Polymer density changes were calculated from mass and vo- = lume measurements. Microscopic examinations at 20x, 50x, and 100x magnifications showed no color or morphological changes of the polymers for both fluids. For all polymers, volume increases in HLP at a maximum of 0.71 % while density decreases at a maximum of 0.89 %. These effects are more pronounced with higher initial polymer density. The PEEK derivatives therefore show the lowest swelling. The swelling of both filled and unfilled polymers is almost the same. Since the fluid’s density is lower than that of the polymers, HLP46 is likely stored in the polymer matrix. Over the 42-day measurement period, a decreasing gradient is observed, but saturation is not yet reached. The hardness increases by maximum approx. 1.3 % due to fluid storage. In HEES46, a maximum volume decrease of 0.66 % and a maximum density increase of 0.33 % are observed. Thereby PEEK shows the least shrinkage. PEEK filled with 30 % carbon fibers shrinks slightly less and has a lower density increase than unfilled PEEK. This suggests additives are released or the polymer itself dissolves. Despite the opposite change in volume, there is also a hardness increase of a maximum of about 1.6 % in HEES46. 3.2 Tribometer test results Based on the insertion tests and wear coefficients from the Table 2, PEEK and PAI in HLP46 were selected for tribometer testing. Before measuring friction and wear, the influence of surface treatment was investigated for natural PEEK, as shown in Figure 7. Before the tribometer tests, a run-in process was conducted until the roughness of the contact partners stabilized. In all experiments material transfer to the mating surface was observable. For a grinded pairing (1), the Science and Research 9 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 external cooling heat sink momentum support with force sensor air bearing normal force application + force sensor drive test fluid chamber temperature sensor angle compensation stator disk rotor slipper holder abstracted slipper conical press fitting F U n U T U speed sensor Figure 4: Structure of the “Slipper-on-Disk” Tribometer Science and Research 10 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 81 82 83 84 H [Micro-Shore D] -0.2 0 0.2 0.4 0.6 0.8 Volume Change [%] -1 -0.8 -0.6 -0.4 -0.2 0 0.2 Density Change [%] 0 10 20 30 40 Duration of Insertion [days] 81 82 83 84 H [Micro-Shore D] 0 10 20 30 40 Duration of Insertion [days] -0.2 0 0.2 0.4 0.6 0.8 Volume Change [%] 0 10 20 30 40 Duration of Insertion [days] -1 -0.8 -0.6 -0.4 -0.2 0 0.2 Density Change [%] Figure 5: Insertion test results in HLP46 mineral oil 81 82 83 84 H [Micro-Shore D] -0.8 -0.6 -0.4 -0.2 0 0.2 Volume Change [%] -0.4 -0.2 0 0.2 0.4 Density Change [%] 0 10 20 30 40 Duration of Insertion [days] 81 82 83 84 H [Micro-Shore D] 0 10 20 30 40 Duration of Insertion [days] -0.8 -0.6 -0.4 -0.2 0 0.2 Volume Change [%] 0 10 20 30 40 Duration of Insertion [days] -0.4 -0.2 0 0.2 0.4 Density Change [%] Figure 6: Insertion test results in HEES46 ester oil Ra=0.1514 Rz=1.0037 Ra=0.1546 Rz=0.8069 x50 x50 Disk grinded Slipper grinded (1) Disk grinded Slipper lapped (2) Ra=0.2671 Rz=1.2016 Ra=0.1322 Rz=0.7042 x50 x50 x50 x50 x50 x50 Ra=0.0868 Rz=0.4388 Ra=0.1763 Rz=0.9699 Ra=0.0655 Rz=0.3432 Ra=0.1398 Rz=0.7035 Disk lapped Slipper lapped Disk lapped Slipper grinded (3) (4) (1) (2) (3) (4) 0 2 4 6 8 10 K W [mm 3 / Nm] 10 8 6 4 2 0 x10 -8 K W [mm 3 / Nm] Figure 7: Wear depending on the surface treatment at 1 MPa and 5 m/ s contact overheats, resulting in an increased wear coefficient. In this test series, low roughness parameters and anisotropy due to lapping can reduce the wear coefficient by up to 81.8 % compared with grinded surfaces. Thereby a better finish of the harder mating surface has a greater impact. Figure 8 shows measurements of frictional coefficients for lapped contact partners. All polymers exhibit a typical Stribeck-curve, indicating significant hydrodynamics in the test. This results in very low friction at high speeds, with little variation between the polymers. However, even at low relative speeds, the friction coefficients are much lower than in Table 2. Fiber reinforcement in PEEK-CF30 slightly increases friction compared to natural PEEK in contrast to the dry “pin-on-disk” test, while the lowest friction values are achieved with graphite and PTFE reinforcement in PEEK-HPV. All polymers show decreasing friction with increasing load, which appears to follow a saturation curve. Figure 9 shows the wear coefficients of the polymers, while Figure 10 shows the corresponding slipper surfaces. Sorted by contact pressure, the tests were conducted in the order: 1.5, 0.5, 2.5, 1, 2, and 3 MPa. Due to the preceded run-in and the given sequence of tests, in Figure 9, the hydrodynamic effect also visible in Figure 8 is recognizable. For all polymers, the wear coefficient significantly decreases up to a relative speed of 7.5 m/ s and then remains roughly constant. PEEK natural has the lowest wear coefficient. For all PEEK polymers, the wear coefficient are two orders of magnitude better, than in Table 2. The surfaces show no signs of adhesion or overheating, with even abrasion traces. Also, on the disks surface no signs of overheating are recogniz- Science and Research 11 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 PEEK PEEK-HPV PEEK-CF30 PAI x x x x 0.023 0.017 0.025 0.022 x 0.139 x 0.149 x 0.086 x 0.117 0.5 0.5 p C , T=40°C Figure 8: Frictional coefficient depending on relative velocity and contact pressure x10 -8 x10 -8 x10 -8 x10 -8 PEEK PEEK-HPV PEEK-CF30 PAI 7.02·10 -9 8.01·10 -9 8.13·10 -9 3.59·10 -8 45 31.25 20 11.25 5 1.25 15 12.5 10 7.5 5 2.5 3 2.5 2 1.5 1 0.5 pv [MPa·m/ s] v r [m/ s] p C [MPa] 45 31.25 20 11.25 5 1.25 15 12.5 10 7.5 5 2.5 3 2.5 2 1.5 1 0.5 pv [MPa·m/ s] v r [m/ s] p C [MPa] Figure 9: Wear coefficients depending on the pv-product 4.2 Tribometer tests With the aid of the limiting pv-product of minimum 150 MPa m/ s in short term, with adequate cooling, and the wear coefficients of down to 8.01·10 -9 mm 3 / Nm, it can be shown, that polymers can generally be used in oil-hydraulic tribological contacts. The measured frictional coefficients of the selected polymers are up to 38.5 % lower for lowand up to 46.8 % lower for high speeds, than those of special brass alloys, which can significantly contribute to increase efficiency. However, the measured wear coefficients suggest that without changes to contact geometry and hydrostatic relief, long-term lifetime issues may arise, as they are approx. a factor 81 worse than those of special brass. Nevertheless, the results indicate that improved lubrication and cooling through hydrostatic relief could further reduce wear coefficients, despite hydrostatic pressure increases the absolute surface stress, influencing wear mechanisms. Li demonstrates this for an “pin-on-disc” tribometer and an axial piston pump under water lubrication [17]. To fully evaluate the potential of polymers in oil-hydraulic contacts, thus wear must be tested in a hydrostatically relieved contact. Due to the high contact deformation, this must be conducted on a real, geometrically adapted slipper to avoid local excess pressure. 5 Conclusion and outlook This publication shows that PEEK, PAI, and PI are suitable for use in mineral oil based on insertion tests. The Science and Research 12 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 able, despite significant material transfer. In fiber-reinforced variants, a graphite solid lubricant layer is visible, which breaks down at higher loads without increasing the wear coefficient. To determine the limiting pv-product, the pressure was increased by 0.25 MPa steps for 30 seconds each, starting from 3 MPa at 15 m/ s. In a dry test at 0.5 MPa and 2.5 m/ s, this duration was sufficient to wear the slippers down to their holder. In contrast, in the oil-lubricated tests, up to 150 MPa·m/ s no significantly increased wear could be detected. It is therefore concluded that the polymers can withstand this load, at least in the short term, with sufficient cooling. 4 Discussion 4.1 Insertion tests All tested polymers swell slightly under HLP46, with PEEK proving to be particularly resistant. This must be considered when designing the hydrostatic relief, as even less than one percent swelling is significant for precise contact balancing. The swelling saturation point still needs determination. An approach for application is to use polymers, that have already reached swelling saturation for precise contact manufacturing. Under these conditions, all tested polymers seem suitable for HLP46. However, since additives may be released or the polymer itself may dissolve in HEES46, more pronounced property changes are expected despite similar hardness changes. Long-term degradation under HEES46 should be monitored, as use in this fluid might be problematic. Ra=0.1240 Rz=0.6242 Ra=0.1359 Rz=0.7097 Ra=0.1376 Rz=0.7070 Ra=0.1300 Rz=0.6649 Ra=0.1417 Rz=0.6961 Ra=0.1802 Rz=1.0325 Ra=0.1259 Rz=0.6869 Ra=0.1337 Rz=0.7075 Ra=0.1362 Rz=0.7105 Ra=0.1760 Rz=0.8365 Ra=0.1587 Rz=0.8206 Ra=0.1285 Rz=0.6124 pv- Stage Polymer after finish PEEK PEEK−CF30 PEEK−HPV PAI x50 x50 x50 x50 x50 x50 x50 x50 x50 x50 x50 x50 Figure 10: Surface structures during wear tests limiting pv-product demonstrates that these polymers can withstand the loads in oil-hydraulic tribological contacts. They have significantly lower friction than special brass but higher wear. Further measurements on hydrostatically relieved contacts are needed for a final evaluation of their potential. Acknowledgement This publication was created as part of the research project “Validation of a simulation methodology for the design of additively manufactured polymer slippers for oil-hydraulic piston machines” (funding code SCHM 3289/ 18-1) in cooperation with the Exzellenzcluster 2186 “The Fuel Science Center”, both funded by the German Research Foundation DFG. The authors would like to thank the DFG for this support. References [1] Findeisen D., Helduser S., Ölhydraulik Handbuch der hydraulischen Antriebe und Steuerungen, 6 th Edition, Springer, Berlin, 2015, ISBN: 978-3-642-54908-3 [2] Wegner S., Experimental and simulative investigation of the cylinder block/ valve plate contact in axial piston machines, Doctoral Thesis, RWTH Aachen, Shaker, Reihe Fluidtechnik. D108, 2020, DOI: 10.18154/ RWTH-2021- 03796 [3] Gärtner M., Verlustanalyse am Kolben-Buchse-Kontakt von Axialkolbenpumpen in Schrägscheibenbauweise, Doctoral Thesis, RWTH Aachen, Shaker, Reihe Fluidtechnik. D97, 2019, ISBN: 978-3-8440-7182-5 [4] Pietrzyk T., Entwicklung einer Hochdrehzahl-Innenzahnradpumpe für die Elektrifizierung mobiler Anwendungen am Beispiel einer autarken dezentralen elektro-hydraulischen Achse, Doctoral Thesis, RWTH Aachen, Shaker, Reihe Fluidtechnik. D109, 2021, DOI: 10.18154/ RWTH- 2022-00767 [5] Czichos H., Habig K.-H., Tribologie-Handbuch Tribometrie, Tribomaterialien, Tribotechnik, 5 th Edition, Springer Vieweg, Berlin, 2020, ISBN: 978-3-658-29483-0 [6] Ivantysynova M., Ivantysyn J. Hydrostatic Pumps and Motors: Principles, Design, Performance, Modeling, Analysis, Control and Testing, 1 st Edition, Tech Books International, 2003, ISBN: 81-88305-08-1 [7] Schlegel F., Merkel A., Schmitz K., Experimental and Simulative Investigation of a Partially Hydrostatic Relieved Contact in Variable Speed Axial Piston Machines, Tribologie und Schmierungstechnik, 70/ 4-5, 2023, DOI: 10.24053/ TuS-2023-0023 [8] Manring N. D., Fluid Power Pumps & Motors Analysis, Design, and Control, 1 st Edition, McGraw-Hill Education, New York, 2013, ISBN: 978-0-07-181220-7 [9] Reetz B., Münch T., Challenges of novel lead-free Alloys in Hydraulics, Proceedings of the 12 th International Fluid Power Conference Vol1-Sym1, pp.17-25, Technische Universität Dresden, Dresden, 2020, DOI: 10.25368/ 2020.6 [10] Holzer A., Reetz B., Münch T., Schmitz K., Experimentelle Untersuchung zum Einlaufverhalten bleifreier Sondermessinglegierungen, Tribologie und Schmierungstechnik, 69/ 22-1, 2022, DOI: 10.24053/ TuS-2022-0003 [11] Blanke H.-J. et al., Expert Praxislexikon Tribologie Plus: 2010 Begriffe für Studium und Beruf, 2 nd Edition, Expert Verlag, 2000, ISBN: 3-8169-06915 [12] George S. C. et. al, Tribology of polymers, polymer composites and polymer nanocomposites, Elsevier Series in Tribology and Surface Engineering, 1 st Edition Elsevier, Amsterdam, 2023, ISBN: 978-0-323-90748-4 [13] Rokala M., Analysis of Slipper Structures in Water Hydraulic Axial Piston Pumps, Doctoral Thesis, Tampere University of Technology, Pub. 1055, 2012 [14] Schoemaker F., Murrenhoff H., Piston Slippers for Robust Water Hydraulic Pumps, Proceedings of the 11 th International Fluid Power Conference Vol1, pp.236-247, RWTH Aachen University, Aachen, 2018, DOI: 10.18154/ RWTH-2018-224460 [15] Li D. et al, Failure analysis on the loose closure of the slipper ball-socket pair in a water hydraulic axial piston pump, Engineering Failure Analysis, Vol. 155/ 107718, 2024, DOI: 10.1016/ j.engfailanal.2023.107718 [16] Li D. et al, The Difference in Tribological Characteristics between CFRPEEK and Stainless Steel under Water Lubrication in Friction Testing Machine and Axial Piston Pump, Lubricants, 11(4): 158, 2023, DOI: 10.3390/ lubricants11040158 [17] Guan Z. et al., Friction and Wear characteristics of CF/ PEEK against 431 stainless steel under high hydrostatic pressure water lubrication, Materials & Design, Vol. 196/ 109057, 2020, DOI: 10.1016/ j.matdes.2020.109057 [18] Nie S., Lou F., Yin F., Tribological Performance of CF- PEEK Sliding against 17-4PH Stainless Steel with Various Cermet Coatings for Water Hydraulic Piston Pump Application, Coatings, Vol. 9(7)/ 436, 2019, DOI: 10.3390/ coating9070436 [19-24] Mitsubishi Chemical Advanced Material Group, Product Datasheet of: 19 - Ketron 1000 PEEK, rev. 14.02.2024; 20 - Ketron CA30 PEEK, rev. 25.06.2024; 21 - Ketron HPV PEEK, rev. 22.11.2023; 22 - Duratron T4503 PAI, rev. 08.05.2024; 23 - Duratron DF7040G PI, rev. 19.01.2024; 24 - Ertacetal H-TF POM-H, rev. 25.06.2024 Science and Research 13 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0002 1 Introduction Electrical energy from wind turbines (WT) already plays a crucial role in the energy supply of Germany. Energy supply from renewable sources like wind is planned to increase in the nearer future for Germany and the European union [1-3]. A key aspect for energy providers is the Levelized Cost of Electricity (LCOE). A significant share of the overall cost of WTs stems from maintenance and servicing [4, 5]. The main bearing is one key component of a WT. Up to 30 % of WT main bearings experience failure during their planned service life [6-8]. Science and Research 14 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 Influence of wear and manufacturing inaccuracies on the performance of a conical plain bearing main bearings for wind turbines Jan Euler, Georg Jacobs, Timm Jakobs, Thomas Decker, Noah Smeets, Julian Röder* submitted: 19.09.2024 accepted: 6.03.2025 (peer review) Presented at GfT Conference 2024 * Jan Euler, M. Sc. Orcid-ID: https: / / orcid.org/ 0000-0001-9293-3219 Prof. Dr.-Ing. Georg Jacobs Orcid-ID: https: / / orcid.org/ 0009-0009-7683-6350 Timm Jakobs, M. Sc. Thomas Decker, M. Sc. Orcid-ID: https: / / orcid.org/ 0000-0002-3296-7166 Noah Smeets, B. Sc. Julian Röder, M. Sc. Orcid-ID: https: / / orcid.org/ 0000-0002-8701-6384 Chair for Wind Power Drives der RWTH Aachen University. Electrical energy harvested by wind turbines already constitutes the largest portion of the current electricity mix in Germany. With increasing penetration of wind energy in the energy sector, wind turbine availability and reliability become more important during turbine design. One component which is prone to failure is the wind turbine’s main bearing. A damaged main bearing results in long downtimes and costly repairs as the drivetrain needs to be dismantled using large cranes or specialised crane vessels for offshore turbines. One possible remedy is the use of plain bearings as main bearings instead of the commonly used rolling element bearings. Plain bearing main bearings can be designed with segments and thus potentially repaired up-tower without the costly dismantling of the drivetrain. Therefore, they can drastically reduce repair costs in case of bearing failure. To this end the novel FlexPad bearing design was developed at the CWD. The general applicability of the FlexPad design was shown in previous publications. During wind turbine operation, which is influenced by the intermittent weather and wind conditions and includes start-stop events, wear is a crucial aspect that needs to be accounted for during the design stage of plain bearings. Until now, there is no knowledge about the influence of wear on the profiling and performance during production loads of the FlexPad. Kurzfassung This paper presents a method to quantify the influence of wear on the contours profiling. The wear’s effect on the segment profiling is assessed by means of surface measurements of the FlexPad bearing. The complex conical geometry necessitates a novel measurement and post processing approach. Using this new approach, the FlexPad contour is evaluated pre and post testing. Based on the measurements the influence of wear and thus the changed contours on the bearing´s performance is evaluated by means of multi body elastohydrodynamic simulations. Moreover, the measured contours profiles are compared to an ideal profile, evaluating the influence of possible manufacturing inaccuracies. The investigated profiles of factory-new segments showed good conformity with the nominal profile, but due to their waviness there is a less favourable pressure distribution. Run-in segments show slight wear and smoothing. Both have a positive effect on the pressure distribution in simulative tests. Schlüsselwörter main bearing, drivetrain, FlexPad bearing, surface measurement, elastohydrodynamic simulations, simulative tests Currently all commercially available WT use rolling bearings as main bearings [9]. The replacement of a faulty rolling element main bearing is especially costly as the rotor needs to be dismounted and the WT drivetrain to be disassembled. One possible approach to address this challenge is the use of plain bearings as WT main bearings. Plain bearings can be designed with segmented sliding surfaces. In case of failure sliding segments can potentially be exchanged individually. This would result in reduced maintenance costs, as the segment exchange could happen up-tower without the use for expensive cranes and the overall downtime would be reduced. Plain bearings as WT main bearings are currently subject to research and development but are not used commercially to date. One such plain bearing design is the FlexPad main bearing which was designed and tested experimentally on different scales at the CWD [10-16]. The Flex- Pad is a double conical plain bearing design with a segmented outer sliding surface which is connected to the housing by a flexible arm structure (see Figure 1). The flexibility allows for the sliding segments to follow the deflection of the shaft which reduces edge wear and improves load distribution. In addition to the macro-geometry design (see Figure 1) a converging gap between shaft and bearing is essential for the hydrodynamic operation of a plain bearing. The converging gap leads to pressure build-up when oil is forced through by the rotation of the shaft. Radial journal bearings achieve this converging gap with an eccentricity between shaft and journal. Segmented bearings can also utilize a surface profile, e.g. a wedge-shaped inlet surface machined on the segment (see Figure 2 right grey areas). The shape of this surface profiling influences the hydrodynamic performance of a plain bearing and improves to the load carrying capacity as well as reducing the frictional losses during the bearing’s operation. The shape was addressed in various studies starting with the simple Rayleigh step from 1918 to 3D topologies in recent studies [18-23]. Naturally, abrasive wear influences the bearings profiling and thus has influence on its hydrodynamic performance. As with most common segmented plain bearings the FlexPad also utilizes profiling on its sliding surfaces to improve performance [24]. The design for the initial profiling was done by hand. To the authors knowledge there are no studies focussing on the effect of profiling for conical plain bearings except the works of Jakobs et al. [24]. A deeper understanding of the influence of profiling on the performance of conical plain bearings is therefore necessary. The FlexPad sliding surface contour was designed with a simple profile. The profile of the contour is shown in Figure 2. Figure 2 left shows the depth of removed material from the surface of the ideal circular surface of a slice of the segment. Denoted is the depth of removed material correspon- Science and Research 15 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 Pad area parallel to shaft Areas of retraction forming the profile of the pad Outlet wedge Inlet wedge Direction of shaft rotation and fluid transport over the segment Profile function -15 15 Figure 2: Target profile of the investigated FlexPad design in 2D (left), target profile in 3D (right) Figure 1: Original FlexPad design (1); Schematic depiction of the FlexPad’s flexibility (2) [17] (1) (2) The contour follows a circular progression which exceeds the depth of the segment profile significantly in scale. Furthermore, the measured section does not follow a circular path but cuts straight across the segment (see Figure 3, d). The measured contour also follows a curvature (see Figure 3 e) with only the outer chamfers of the profile shown in Figure 2 being discernible. To determine the machined profile of the segment the underlying curvature must be removed through post processing. As the FlexPad has a conical shape one could assume the measured contour follows a circle. This, however, is not the case. The macro geometry of the measurement is analogous to a conic section. As the cutting plane is roughly parallel to the opposing generating line of the cone, the measured macro geometry can be approximated with a parabola. A parabola with its apex and the coordinate centre can mathematically be described as: 1 For a conic section the parameter a can be defined as 2 The length AF −− refers to the distance between the apex and the focal point of the parabola. The focal point of the parabola stemming from a conic section is at the intersection between the cutting plane and the cone’s centre line. Through the breadth of the measured contour, the axial position of the measurement along the segment and thus its total position can be determined. The Flex- Pad segments are measured with a 90° angle respective to the backside of the segment. The cone angle of the FlexPad V2 is 46.7° [12]. Using this knowledge, the focal point and its distance to the apex can be determined, as shown in Figure 4. With the function of the parabola known for each measurement, its value can be deducted from the measured contour. This results in a profile measurement similar to the target profile shown in Figure 2. 2.2 Measurement Results To establish a baseline for the observed wear and damage, the profile of a segment which has not been used in test bench trials is measured. The results are shown in Figure 5. Depicted is the following: 1. The target profile. 2. The contour as measured and adjusted by the parabola function, which was determined through the measurement position. The measured contour shows good agreement with the target contour. Between ten and fifteen degrees of centre ( ) = = 1 4 Science and Research 16 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 ding to the angular position on the sliding surface. The profiling is symmetrical to allow for usage of segments for both rotor side (RS) and generator side (GS). The profile is most pronounced at the edges of the sliding segment, where the most material is removed to further reduce the possibility of edge wear. Towards the centre the profile follows a slight ramp to induce the formation of a convergent lubrication gap. In the centre of the segment the profile forms a plateau and no material is removed. This allows for large load bearing surface. The sliding surface of the FlexPad’s sliding segments consist of a thermally sprayed coating of cobalt-chromium alloy with integrated hexagonal bohrnitrite. The surface material was especially chosen for its wear resistance and low friction coefficient [16, 25]. Nevertheless, wear can still be expected to occur during the bearing’s lifetime, due to the special load conditions experienced by a WT’s main bearing (e.g. start-stop operation under the influence of high radial load) where mixed friction operation cannot be avoided completely. In this study the wear driven profile change between factory new and tested segments of the FlexPad bearing is investigated. To this end a simple measurement approach is presented to determine the contour and its profile for conical sliding surfaces. The profile of the measured contours of worn and unworn segments are compared to the target profiling of the contour. The influence of wear and manufacturing inaccuracies on the FlexPad profile and on its performance is assessed by means of a multi-body elasto-hydrodynamic (MB-EHD) simulation. 2 Contour Measurements To assess the effect of wear and manufacturing inaccuracies on the bearing’s performance, the profile of tested segments needs to be determined. Currently, there is no standardised method for the contour and profile measurement of conical plain bearings. The conical shape results in the contour dominating the measurement. An assessment of the profile is not possible without postprocessing. Therefore, a novel approach is proposed in this study. 2.1 Measurement Method To measure the contour of the FlexPad segments a Hommel T8000 (see Figure 3, c) is used. The measurement system has a resolution of 0.1 μm. The contour is determined by a tactile measurement along the Xand Ycoordinates of the segment surface. The individual segments are measured in 10 distinct radial measurement positions as shown in Figure 3 b. The measurement positions are identical for all segments and are determined by individual positioning slots in the measurement bracket (see Figure 3, a and b). For common radial plain bearings a measurement like this results in a directly visible profile of the measured plain bearing [26]. The approach however is not suitable for the FlexPad geometry. the desired chamfers are distinctly visible. In the centre (± 5 °) no material is removed from the original circular sliding surface for the factory new segments. Between the chamfers and the centre inlet and outlet wedges are visible, starting at around ± 6.5 ° of centre. Unlike the target contour, the measured contour features a slight waviness. The influence of the minimal manufacturing inaccuracy on the bearing’s performance is further analysed in chapter 3. Some segments on the rotor side show clear indications for three body abrasion (see Figure 7, right). The abrasion groove follows a circular pattern. Therefore, it was most likely caused by metallic debris dragged around for one or more revolutions causing abrasive wear. The abrasion is present in almost all RS segments and becomes more pronounced towards the centre of the loaded zone for production loads and start/ stop procedures. The circular abrasion can also be documented with the contour measurements (see Figure 8). The parabola sub- Science and Research 17 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 a) b) c) d) e) measurment positions direction of measurment μ Figure 3: Bracket for segment measurement (a and b), contour measurement system (c), linear measurement progression (d), Measurement without post processing (e) 46.7 ° Measurement position Focal point 90 ° Figure 4: Geometric determination of parabola focal point centre two distinct collapses are visible. This is caused by the measurement progression path cutting through the groove two times, as the circular arc of the groove is Science and Research 18 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 traction derived from the measurement position yields expected results. The chamfers and inlet/ outlet wedges are relatively unfazed by the abrasion. In the segment start/ stop wear groovelike abrasion Figure 6: Lower RS bearing halve top view (left), segment in three o’clock position with circular groovelike abrasion (right) Rotation Figure 5: Contour for an untested segment in comparison with the target contour Rotation Figure 7: Contour measurements for a rotor side segment in abrasion grove height traversed. For these contours no following MB-EHD simulation is performed as the abrasion is only localized in a small segment area. The FlexPad prototype was subject to intense testing of startand stop procedures. The amount of start-ups corresponds to the amount a WT main bearing would see during its predicted service live [10]. The segments in the lower halve of the RS cone show signs of wear and some groovelike abrasion in accordance with the remaining segments. In Figure 8 the measured contour of a tested RS segment is shown in comparison to the target contour. The tested segment was mounted in the bottom position of the RS as shown in Figure 6. Again, the profile determined via parabola subtraction is shown. The measurement position is in the segment centre where significant wear is visible. The measured profile is visibly less wavy than the contour of the untested segment, see Figure 5. The contour seemingly follows the same progression as for the untested segment. However, the inlet wedge progresses to the segment centre instead of stopping at ± 6.5 °. As the contour measurement is aligned to 0 mm at its highest point, the total depth of material loss cannot be determined using this method. Overall the contour shows little signs of wear and material removal. The visible smoothening of the segment however, warrants further investigation in future studies. 3 Simulation Results and Discussion Contours and their profiling have a large influence on the bearing’s performance. To assess the influence of the worn profile for the FlexPad bearing, MB-EHD simulations were performed using the determined contours. The simulation setup is identical to previous studies [10, 12, 24, 27, 28]. The contour profile was integrated with a grid size of 4 mm. Investigated was the bearing’s performance under stationary production load conditions. For all simulations one profile was applied to all segments and over the whole segment length. For the target profile this is clearly suitable as all segments were designed with this profiling in mind. For the measured contour of factory new segments this already constitutes a compromise as not all measured contours were 100 % identical. However, all showed the same general profile progression. Therefore, a representative profile was chosen from a centre measurement position (see, Figure 5). Contours that documented the groovelike abrasion were therefore unsuitable for simulation as the area could not sensibly be extended over the whole segment. For worn contours from segments of the bottom RS position a representative profile was chosen and applied for all segments (see, Figure 9). The bottom centre on the RS would typically experience the most amount of load during production load conditions and start/ stop procedures. This is due to the weight of the Rotor. The compromise, to use this contour a representative for the worn FlexPad, was deemed to be acceptable as a worst-case investigation. Future studies should further investigate the influence of 3D contour profiling for individual segments and non-uniform contour profiling for all segments on the bearing’s performance. The results of the simulative investigation on the bearings performance for different profiles are listed in Table 1. Listed are the values for the maximum hydrodynamic pressure, the maximum average segment pressure, the total amount of friction loss, the percentage of pressured area and the amount of solid body contact force relative to the sum of all forces. The target profile achieves the lowest maximum pressure of 17.9 MPa. The factory new profile results in a 65 % increased maximum pressure relative to the target contour. The worn contour has a significantly lower maximum pressure relative to the target profile (11 % increase). The increased maximum pressures for the factory new profile are a result of the overall wavi- Science and Research 19 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 Rotation Figure 8: Contour of tested segment from the bottom RS position The worn profile shows further reduced friction loss, due to the slightly reduced pressured area and lack of solid body friction. Bearings featuring the factory new profile demonstrate the lowest amount of hydrodynamically pressured area and are the only sample for which solid body contact would occur. In Figure 9 the difference in hydrodynamic pressure distribution corresponding to the difference in oil film height for worn and factory new contours is shown. For the most highly pressured segment (see Figure 9, highlighted segments) it can be seen, that the waviness of factory new contours results in the same waviness for the lubrication gap. The sudden fluctuations in the height of the lubrication gap result in a smaller loaded zone for the induvial segments and thus higher maximum pressu- Science and Research 20 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 ness and slight manufacturing inaccuracies of the nonidealized profile. The improved performance of the worn profile is due to the overall less wavy surface which features an elongated inlet wedge. An elongated inlet wedge was already demonstrated to yield positive effects on pressure distribution for the FlexPad [24]. Both waviness reduction and the elongated inlet wedge are a result of the running in process, which is common for plain bearings. The maximum average segment pressure is comparable for all profiles, as the overall load distribution in the FlexPad bearing is unaffected by the profile. The worn profile shows a slightly reduced maximum average segment pressure. This is caused by a slightly improved force distribution among segments. Simulated bearings with the factory new profiles demonstrate the highest friction loss, caused by the solid body contact. Target contour Factory new contour Worn contour Maximum hydrodynamic pressure [MPa] 17.9 29.5 19.8 Maximum average segment pressure [MPa] 2.08 2.12 2.01 Friction loss [W] 407 427 387 Pressured area*) [%] 47.8 44.4 46.7 Solid body contact**) [%] 0 0.2 0 *) area of the bearing surface with hydrodynamic pressure above 1 Pa **) amount of force carried through solid body contact relative to the sum of all forces Table 1: Simulation result for different contours under production load conditions worn contour worn contour factory new factory new Oil film height [μm] Figure 9: Oil film height (left) and hydrodynamic pressure (right) for both a FlexPad bearing with factory new and worn contours (shown is the RS in the front), segment with highest pressure highlighted in black res. The area of maximum pressure is also the area for which solid body contact occurs. For the investigated load conditions the target profile performed best. Factory new profiles displayed waviness which was disadvantageous regarding pressure distribution and thus maximum pressures. Maximum pressures are significantly increased for the measured contour, caused by the waviness of the profile. The increased maximum pressures could lead to damage in the running in process and should be kept in mind during bearing design. The worn contour profile showed clear improvement compared to factory new contour profiles due to running in. Worn profiles displayed no waviness and featured advantageously elongated inlet wedges. For the worn profiles the bearing achieves a performance comparable to the idealised target profile. The wear experienced by the FlexPad during test procedure is therefore not critical. The FlexPad segments contour profiling can be expected to remain functional during the service life of its wind turbine, should no extraordinary events occur. 4 Conclusion Plain bearings play an ever-greater role in drivetrains of WTs. Current industry and academic research explore the feasibility of plain bearings as WT main bearings. Modern plain bearings feature sophisticated surface profiles to improve performance and guaranty safe operation. In this study the influence of wear on the profile of the novel conical plain bearing main bearing concept FlexPad was determined, using 2D contour measurements with post processing. Investigated were factory new segments and worn segments from a highly loaded position within the bearing. The determined profiles were compared to the target profiles in their progression. Further their effect on the bearing’s performance during production load conditions was investigated by means of simulations. The progression of the factory new profiles showed a very good agreement with the target profiles Profiles of worn segments showed expected signs of running in i.e. smoothing and elongated inlet wedges. During simulations of production load conditions, the bearings with factory new profiles demonstrated significantly higher maximum hydrodynamic pressures than the idealised target profile. The worsened behaviour is most likely caused by the profile’s waviness and slight manufacturing inaccuracies. For the worn profiles the bearing’s performance was comparable to results for the idealised target profiles with slight improvements in friction loss and maximum hydrodynamic pressures. The measured contour showed little wear, which improved the bearings performance rather than worsening it. The FlexPad profile therefore demonstrated good wear resistance for the approximated service live of a wind turbine, as more than 10.000 start stop procedures were performed. Future studies will further combine the presented method with the 3D profile methods investigated by Jakobs et al. [24] to achieve an improved understanding of the effect of homogeneously worn contour profiles. The obtained measurements will also be used to assess the influence of surface roughness on breakaway torque and overall performance, i.e. maximum pressures, friction loss and load distribution. References [1] Climate Action: 2050 long-term strategy. URL https: / / climate.ec.europa.eu/ eu-action/ climate-strategies-targets/ 2050-long-term-strategy_en. - Aktualisierungsdatum: 2023-07-20 - Überprüfungsdatum 2023-07-24 [2] German Bundestag: Zweites Gesetz zur Änderung des Windenergie-auf-See-Gesetzes und anderer Vorschriften (in Kraft getr. am 2022) (2022) [3] European Parliament: Amendments adopted by the European Parliament on the Renewable Energy Directive (in Kraft getr. am 2022) (2022) [4] Tyler Stehly, Patrick Duffy: 2021 Cost of Wind Energy Review [5] Tyler Stehly, Philipp Beiter, and Patrick Duffy: 2019 Cost of Wind Energy Review [6] Hart, Edward; Turnbull, Alan; Feuchtwang, Julian; Mcmillan, David; Golysheva, Evgenia; Elliott, Robin: Wind turbine main-bearing loading and wind field characteristics. In: Wind Energy 22 (2019), Nr. 11, S. 1534- 1547 [7] Edward Hart, Kaiya Raby, Jonathan Keller, Shawn Sheng, Hui Long, James Carroll, James Brasseur, Fraser Tough: Main Bearing Replacement and Damage-A Field Data Study on 15 Gigawatts of Wind Energy Capacity [8] Hart, Edward; Stock, Adam; Elderfield, George; Elliott, Robin; Brasseur, James; Keller, Jonathan ; Guo, Yi ; Song, Wooyong: Impacts of wind field characteristics and nonsteady deterministic wind events on time-varying mainbearing loads. In: Wind Energy Science 7 (2022), Nr. 3, S. 1209-1226 [9] Nejad, Amir R.; Keller, Jonathan; Guo, Yi; Sheng, Shawn; Polinder, Henk; Watson, Simon; Dong, Jianning; Qin, Zian; Ebrahimi, Amir; Schelenz, Ralf; Gutiérrez Guzmán, Francisco; Cornel, Daniel; Golafshan, Reza; Blockmans, Bart; Bosmans, Jelle; Pluymers, Bert; Carroll, James; Koukoura, Sofia; Hart, Edward; Mcdonald, Alasdair; Natarajan, Anand; Torsvik, Jone; Moghadam, Farid K.; Daems, Pieter-Jan; Verstraeten, Timothy; Peeters, Cédric; Helsen, Jan: Wind turbine drivetrains: state-of-the-art technologies and future development trends. In: Wind Energy Science 7 (2022), Nr. 1, S. 387-411 [10] Abschlussbericht: FlexPad - Entwicklung einer Auslegungsmethodik für ein innovatives Lagerungskonzept, eng.: “Final Report, FlexPad”: Grant no.: 03EE2002B. 2023 [11] Rolink, Amadeus; Jacobs, Georg; Müller, Matthias; Jakobs, Timm; Bosse, Dennis: Investigation of manufacturing-related deviations of the bearing clearance on the performance of a conical plain bearing for the application as main bearing in a wind turbine. In: Journal of Physics: Conference Series 2257 (2022), Nr. 1, S. 12006 Science and Research 21 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 [21] Kalliorinne, Kalle; Almqvist, Andreas: Application of topological optimisation methodology to finitely wide slider bearings operating under incompressible flow. In: Proceedings of the Institution of Mechanical Engineers, Part J: Journal of Engineering Tribology 235 (2021), Nr. 4, S. 698-710 [22] Kalliorinne, Kalle; Larsson, Roland; Almqvist, Andreas: Application of topological optimisation methodology to hydrodynamic thrust bearings. In: Proceedings of the Institution of Mechanical Engineers, Part J: Journal of Engineering Tribology 235 (2021), Nr. 8, S. 1669-1679 [23] Kalliorinne, Kalle; Pérez-Ràfols, Francesc; Fabricius, John; Almqvist, Andreas: Application of topological optimisation methodology to infinitely wide slider bearings operating under compressible flow. In: Proceedings of the Institution of Mechanical Engineers, Part J: Journal of Engineering Tribology 234 (2020), Nr. 7, S. 1035-1050 [24] Jakobs, T.; Jacobs, G.; Euler, J.; Rolink, A.; Röder, J.: Impact of 3D segment profiling on friction losses of plain bearings in wind turbines main bearings. In: Journal of Physics: Conference Series 2767 (2024), Nr. 5, S. 52021 [25] Rolink, Amadeus; Jacobs, Georg; Bosse, Dennis; Schröder, Tim; Bobzin, Kirsten; Öte, Mehmet; Königstein, Tim; Wietheger, Wolfgang: Erprobung von thermisch gespritzten Schichtsystemen in Gleitlagerversuchen (2019) [26] Decker, Thomas; Jacobs, Georg; Graeske, Carsten; Röder, Julian; Lucassen, Mattheüs; Lehmann, Benjamin: Multiscale-simulation method for the wear behaviour of planetary journal bearings in wind turbine gearboxes. In: Journal of Physics: Conference Series 2767 (2024), Nr. 5, S. 52012 [27] Euler, Jan; Jacobs, Georg; Jokobs, Timm; Röder, Julian: Feasibility study on preloaded flexible conical plain bearings as wind turbine main bearings. In: Tribologie und Schmierungstechnik 71 (2024), Nr. 1 [28] Jakobs, T.; Jacobs, G.; Rolink, A.; Euler, J.; Hollas, C.; Röder, J.: Designing Large Segmented Flexible Conical Plain Bearings as Wind Turbine Main Bearings. In: Journal of Physics: Conference Series 2745 (2024), Nr. 1, S. 12022 Science and Research 22 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0003 [12] Rolink, Amadeus; Jacobs, Georg; Pérez, Alex; Bosse, Dennis; Jakobs, Timm: Sensitivity analysis of geometrical design parameters on the performance of conical plain bearings for use as main bearings in wind turbines. In: Journal of Physics: Conference Series 2265 (2022), Nr. 3, S. 32010 [13] Rolink, Amadeus; Jacobs, Georg; Schröder, Tim; Keller, Dennis; Jakobs, Timm; Bosse, Dennis; Lang, Jochen; Knoll, Gunter: Methodology for the systematic design of conical plain bearings for use as main bearings in wind turbines. In: Forschung im Ingenieurwesen 85 (2021), Nr. 2, S. 629-637 [14] Rolink, Amadeus; Schröder, Tim; Jacobs, Georg; Bosse, Dennis; Hölzl, Johannes; Bergmann, Philipp: Feasibility study for the use of hydrodynamic plain bearings with balancing support characteristics as main bearing in wind turbines. In: Journal of Physics: Conference Series 1618 (2020), Nr. 5, S. 52002 [15] Schröder, Tim; Jacobs, Georg; Rolink, Amadeus; BossE, Dennis: “FlexPad” - Innovative conical sliding bearing for the main shaft of wind turbines. In: Journal of Physics: Conference Series 1222 (2019), Nr. 1, S. 12026 [16] Abschlussbericht: Thermisch gespritzte Gleitlagerbeschichtungen für Hauptlager von Windenergieanlagen (WEA) - WEA Triebstrang und Oberflächentechnik, eng.: “Final Report, WEA-GLiTS”: Förderkennzeichen: 03EK3036A [17] Schröder, Tim Niklas: Konisches Gleitlager für die Rotorlagerung einer Windenergieanlage, eng: “Conical Sliding Bearing for the Rotor Main Bearing of a Wind Turbine” (2021) [18] Fesanghary, M. ; Khonsari, M. M.: Topological and shape optimization of thrust bearings for enhanced load-carrying capacity. In: Tribology International 53 (2012), S. 12-21 [19] Kettleborough, C. F.: The Hydrodynamic Pocket Thrust- Bearing. In: Proceedings of the Institution of Mechanical Engineers 170 (1956), Nr. 1, S. 535-544 [20] Rayleigh: I. Notes on the theory of lubrication. In: The London, Edinburgh, and Dublin Philosophical Magazine and Journal of Science 35 (1918), Nr. 205, S. 1-12 Science and Research 23 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 Submit your paper until 31st of May 25 More information and registration www.tae.de/ 50019 25 th International Colloquium Tribology Bridging Science and Industry - Driving a Sustainable Future with Tribology Join Europe`s leading conference on lubrication, friction and wear! Experience 3 intensive days featuring 130 presentations from top experts in research, industry and practice across 5 parallel sessions, attracting over 400 participants from around the globe. Don‘t miss the special 25th anniversary edition — save the date today! Ost昀ldern/ Stuttgart, Germany 27th - 29th January 2026 the characteristics of the running-in process. The contact geometry includes the radii of curvature of the bodies, waviness and the surface roughness and structure. Kinematics and load can include, for example, the rolling motion, the normal and frictional force and the contact duration. Due to the running-in, a change in the coefficient of friction with the variation in speed can be determined in friction force tests. Depending on the minimum speed or lubricant film thickness, a more or less pronounced hysteresis results for the coefficient of friction. Knowledge of the running-in hysteresis can contribute to targeted optimization when designing tribological systems. Science and Research 24 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 1 Introduction, Motivation und Objective Highly stressed rolling-sliding contacts, such as the tooth flank contact, are tribological systems, for short tribosystem [CZIC10]. The properties of the tribological system are determined by the interactions between the contact partners, the lubricant and the ambient medium [POPO10]. For the tribosystem “gear”, the base and counter body can be defined as pinion and wheel, and the intermediate and ambient medium as gear oil and air. Depending on the initial state and the operating conditions, the tribological system changes continuously over time [PREX90, VOLG91, LÖPE15, LOHN16]. The change in tribological contact conditions is most pronounced during the initial loading of the tribosystem. This initial change is referred to as the running-in phase [ASMI92, CZIC10]. On the one hand, local deformation of the active surface takes place during the run-ning-in process (contact mechanics), which can be attributed to a local exceeding of the material strength or yield stress and can be quantified by recording the geometric surface properties. On the other hand, interactions of the triboelements can take place at the atomic and molecular level during runningin, which can be responsible for the formation of boundary layers as a result of adhesion, physisorption and chemisorption, for example [SCHU10, BREC16]. Figure 1-1 shows the variables influencing the running-in process in the rolling contact. The running-in process combines the initial state of the rolling surfaces (after production) with the state during operation. In addition to the kinematics and the lubrication, the contact geometry, the load and the material properties of the surface zone are among the central influences on Elastic-plastic micro contact calculation for large scale, lubricated rolling-sliding contacts Dieter Mevissen, Sebastian Sklenak, Christian Brecher, Thomas Bergs* Presented at GfT Conference 2024 The surface roughness and structure of rolling-sliding contacts have a decisive influence on the running behavior in terms of friction, wear and fatigue. Due to running-in effects, there is a continuous change of the surface roughness during the first load cycles. For this reason, the aim of this work was a calculation method to predict the friction force hysteresis during running-in. The basis of the method is a large-scale micro contact calculation based on the elastic halfspace theory in combination with an analytical stress calculation. The material behavior is modeled elastoplastically, whereby a limiting stress is used as a yield criterion instead of the limiting pressure usually used in the elastic half-space. The influence of the lubricant pressure is implemented by a lubricant boundary condition for the contact calculation in the elastic halfspace according to B OUSSINESQ , L OVE and H ARNETT . Finally, the calculation method is validated by means of friction force tests. For this purpose, S TRIBECK curves were recorded on a disc-on-disc test rig and compared with the calculation results. Keywords Micro contact calculation, half-space theory, Boussinesq, Love, Harnett, rolling-sliding contact, friction force measurement, lubricant boundary condition, macro and micro scale Abstract * Dr.-Ing. Dieter Mevissen Sebastian Sklenak M.Eng. Prof. Dr.-Ing. Christian Brecher Werkzeugmaschinenlabor der RWTH Aachen Campus-Boulevard 30, 52074 Aachen Prof. Dr.-Ing. Thomas Bergs Manufacturing Technology Institute MTI der RWTH Aachen Campus-Boulevard 30, 52074 Aachen This article presents a calculation method for predicting the friction force hysteresis during running-in (Figure1-2). The basis of the method is a large-scale microcontact calculation in the elastic half-space in combination with an analytic stress calculation. The material behavior is modelled elastic-plastic, whereby a limiting stress is used as a yield criterion instead of the limiting pressure usually used in the elastic half-space theory. The influence of the lubricant pressure is implemented by a lubricant boundary condition for the contact calculation in the elastic half-space according to B OUSSINESQ , L OVE and H ARTNETT . Finally, the calculation method is validated by means of friction force tests. For this purpose, S TRIBECK curves were recorded on a two-disk test rig. Each series of measurements started with a speed run-down in order to achieve continuous smoothing of the surface. This was followed by a speed run-up to determine the hysteresis of the friction coefficient with the smoothed surface. The experimental procedure was simulated on the basis of the measured surfaces and finally compared. 2 Elastic-plastic and stress-based micro-contact calculation The main challenge in designing the calculation method is the realization of a large-scale contact calculation with sufficient resolution to capture the surface roughness and structure. Figure 2-1 shows the schematic structure of the calculation method. The starting point of the calculation is the surface topography of the contact bodies Science and Research 25 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 Figure 1-1: Influence factors on running-in in tribological systems © WZL Figure 1-2: Objective and approach © WZL the plastic material behavior on the basis of a yield criterion. In the third step, the elastic stress tensor is converted into a plastic strain tensor using the plasticity model according to PRANDTL-REUSS [LEE12]. Finally, the deformation in the normal direction to the surface is calculated using the semianalytical approach according to JACQ (4 th step) [JACQ02]. A detailed description of the individual calculation steps is summarized in [MEVI21]. Key parameters and factors influencing the calculation result are discussed below. During the stress calculation in the micro-contact, interactions occur between the stress fields of neighboring roughness peaks, see Figure 2-2. To analyze the interactions, patterns with different distances are generated based on the pressure distribution for the reference model, Science and Research 26 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 in the initial state. The calculation method is based on an iterative solution algorithm with four main calculation steps. The iterative solution procedure is necessary due to the pronounced non-linearities in the elastic-plastic micro-contact calculation. The output variable of the calculation is the surface topography in the running-in state with a sufficient size for a standard-compliant roughness evaluation. In the first step, the pressure distribution for the purely elastic contact is calculated (Figure 2-1). The two-dimensional pressure distribution within the contact surface is then converted into a three-dimensional stress state on the basis of AIRY’s stress functions [AIRY62]. The stress tensor for each volume element forms the basis for calculating an equivalent stress and for modelling Figure 2-1: Structure of the Calculation Method © WZL Figure 2-2: Difference of pressure and volume-stress based plasticity modell © WZL in which the pressure distribution is duplicated. This creates a model contact. The maximum pressure of p H = 3139 MPa is the same for all calculations, so that the variants do not differ in the case of a pressure-based yield criterion. Figure 2-2 above shows the influence of the distance between two neighboring point contacts on the hydrostat and the MISES equivalent stress. The results are evaluated at the location of the maximum equivalent stress and normalized to the variant without direct neighbors (infinite distance). As the distance decreases, the hydrostat increases in magnitude and consequently the equivalent stress (deviator) for the contact point in the center decreases. The change is due to the superposition of stress components caused by radiation effects from the neighboring micro-contacts. In addition to the distance, the number of contact points has an influence on the hydrostat and the equivalent stress, see Figure 2-2 below. If the number of direct neighbors is further increased by a cross pattern, the hydrostat increases in magnitude. The increase converges to a limiting value, which is physically to be expected because the influence of highly distant contact points continuously decreases with increasing distance. The maximum reduction of the MISES equivalent stress is approx. 16 % for the examples presented here. In addition to the interaction between individual roughness contacts, knowledge of the permissible evaluation range in the depth direction is a key requirement. In micro-contact calculation, the minimum evaluation depth as a function of the meshing resolution on the surface plays a particularly important role. Figure 2-3 on the left shows the stress component σ x as a function of depth for three different crosslinking resolutions (dy, dz) on the surface (reference model, dx = 0.1 µm). While at greater depths (x > 0.01 mm) the stress curves are identical for the different crosslinking resolutions, a difference is recognizable at shallower depths, with the stress component σx converging towards negative infinity in all cases for x = 0 mm. According to the analytical calculation, this course is not to be expected, as the magnitude of the stress component σx forms its maximum directly at the surface at x = 0 with a magnitude equal to the contact pressure according to H ERTZ p H . It can also be seen that the stress curves reach the amount of H ERTZ pressure at different depths. The higher the meshing resolution, the lower the depth at which the H ERTZ stress is first reached. To quantify the minimum evaluation depth, a convergence analysis is carried out as a function of the surface crosslinking in Figure 2-3 on the right. The minimum evaluation depth is the numerical limit up to which the semi-analytical stress calculation can be evaluated depending on the meshing resolution. The criterion for determining the minimum evaluation depth is the depth at which the normal stress σ x (normal to the surface) assumes the value of the H ERTZ stress. In the case of square meshing, the minimum evaluation depth decreases almost linearly with the meshing resolution at the surface, see Figure 2-3 on the right. For rectangular surface elements, the minimum evaluation depth is determined by the longer edge length, so that an optimization of the calculation time for the two-dimensional pressure calculation with rectangular elements is at the expense of the minimum evaluation depth. To define the resolution dy, dz in the finely meshed contact area by, bz, a convergence study is carried out for the stress and deformation calculation. For the convergence analysis, a contact calculation is also carried out for a section of a rough surface. In the stress calculation, the individual microstress fields are resolved with suffi- Science and Research 27 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 Figure 2-3: Minimum evaluation depth depending on the surface mesh size © WZL re 3-1 shows the solution of the elastic half-space taking into account a lubricant boundary condition. With the lubricant boundary condition, the pressure in the areas without solid contact is not set to zero, but to the value of the lubricant pressure. As the lubricant pressure causes deformation cross influences on the surrounding elements, the right-hand side of the linear system of equations changes for all elements. It is therefore no longer possible to simply delete the rows and columns, as is done with the adhesion boundary condition due to setting the contact pressure to zero. Therefore, the linear system of equations is solved in the first step and then sorted by size. A negative contact pressure exists if the element has to be contracted for contact determination. This is precisely the case for the roughness valleys, which lie behind the roughness peaks and do not come into contact despite the elastic flattening. Finally, the lubricant boundary condition is applied in this step, in which a negative contact pressure is set equal to the lubricant pressure. In the third step, the linear system of equations is separated into a variable and a constant term at the transition from positive to negative contact pressure, see Figure 3-1. The constant term no longer changes in the further elastic contact calculation because the average lubricant pressure in the roughness valleys is specified as constant. In the last step, the linear system of equations is reduced and the constant term on the left-hand side is subtracted from the vector on the right-hand side. It is precisely at this point that the deformation cross influences caused by the lubricant pressure are taken into account by the subtraction when solving the linear system of equations. The solution steps are repeated until no more negative contact pressures occur after the first iteration step. Science and Research 28 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 cient accuracy from a resolution of dy = dz = 0.8 µm, see Figure 2-4 left. To compare different meshing resolutions, the stress depth curves for a contact point in the microcontact are also evaluated. It can be seen that the stress curves converge, whereby deviations can be seen for a resolution of dy = dz = 2.0 µm, particularly at greater depths compared to a finer resolution. Using the same procedure, the local changes in the contact geometry for different meshing resolutions are evaluated in the deformation calculation, see Figure 2-4 on the right. A positive deformation value means that the surface has risen at this point. The amounts of deformation on the surface converge with increasing meshing resolution. If the mesh is too coarse (dy = dz = 2.0 µm), the deformation tends to be calculated too small. From an element size of dy = dz = 0.8 µm, the calculation results for the example contact are comparable. Based on the convergence analysis for the stress and deformation calculation, the high-resolution contact area (by, bz) must therefore be meshed with an element size of at least dy = dz = 0.8 µm. The meshing resolution at the surface results in a minimum evaluation depth of x min = 1.3 µm. 3 Lubricant pressure within elastic half-space theory The elastic half-space offers an efficient way for solving the contact problem of two bodies with arbitrary geometry and is suitable for predicting the geometric surface change in fluid-free contact. In this section, a model for the lubricant pressure in the elastic half-space is presented and a boundary condition for the lubricant volume equilibrium is explained according to [MEVI19]. Figu- Figure 2-4: Convergence Study for Plastic Deformation © WZL In addition to the lubricant boundary condition, the calculation of mixed friction states requires the coupling of the micro-contact calculation based on the elastic half-space theory with the knowledge from the contact calculations of the EHD theory with ideally smooth surfaces. In contrast to the established procedure in the literature, the contact calculation in the elastic half-space is used as the main model, which draws on the results of the EHD contact calculation (auxiliary model) [BART10, SAUE18]. The basic idea of the calculation approach is that a certain volume of lubricant is conveyed into the lubrication gap by the hydrodynamics of the lubrication gap and is distributed evenly in the contact geometry deformed under load. Based on this basic idea, a further boundary condition can be defined for the micro-contact calculation in the elastic half-space during iterative solution finding, which takes into account the balance between the hydrodynamically conveyed lubricant volume and the pocket volume of the roughness valleys under load, see Figure 3-2 center. The boundary condition of the lubricant volume equilibrium is the link between the calculation results of the EHD theory at macro level and the micro-contact calculation in the elastic half-space. In iterative solution finding, the fulfilment of this new equilibrium condition is made possible by the degree of freedom gained in the specification of the lubricant pressure in the area of the roughness valleys. Science and Research 29 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 Figure 3-1: Solving algorithm for Lubricant Boundary Condition for elastic half-space © WZL Figure 3-2: Determination of Lubricant Pressure © WZL intersects the calculated curve beforehand, so that a potential for reducing the Ra value remains with a further reduction in speed. The comparison with the experiment is based on the measured and calculated roughness values at the end of the speed run-down, see Figure 4-1 bottom right. For the increasing normal force, the calculation correctly maps the increasing smoothing. In addition, the degressive course of the smoothing with increasing normal force is captured by the calculation. Overall, the change in the Ra value for all three normal forces is overestimated by the calculation, leaving room for optimizing the elastic-plastic material behaviour and modelling the material distribution. The basis for the calculation of the friction coefficient is the solid contact ratio, which is a central output variable from the large-scale micro-contact calculation. Figure 4-2 shows the change in the solid contact ratio for the different normal forces plotted against the average lubrication pocket height. According to the experiment, an elastic-plastic microcontact calculation (speed downrun) is carried out in the first step. This is followed by a purely elastic contact calculation (speed run-up). The reversal point between speed run-down and run-up depends on the lubricant volume conveyed at minimum speed or the minimum average lubricant film thickness h EHD,min . As the normal force increases, the reversal point shifts to lower solid contact ratios. The maximum possible smoothing is therefore not achieved for the higher normal forces for the operating parameters, see also Figure 4-1. Each time the operating conditions change, the surface roughness is reduced again, resulting in hysteresis for all Science and Research 30 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 4 Prediction of running-in hysteresis in disc-on-disc contact During the determination of S TRIBECK curves by a speed run-down and speed run-up, a hysteresis occurs for the friction coefficient. In the following section, the calculation method is applied to the tests with different normal forces and a prediction of the hysteresis is attempted. The calculation parameters and the material model correspond to the documentation in [MEVI21] in Chapter 7. Figure 4-1 shows the calculated change in the Ra value for the three different normal forces. Due to the elasticplastic contact calculation and the consideration of the lubricant pressure, the Ra value is continuously reduced as the average lubrication pocket height decreases (speed deceleration). The leftmost point of each curve marks the state with a solid contact ratio of ψ load = 100 % and thus the maximum possible smoothing for the respective normal force. Due to the higher total deformation with increasing normal force, the maximum possible smoothing is highest for the largest normal force. The vertical lines in the diagram in Figure 4-1 represent the link between experiment and calculation. The minimum mean lubricant film thickness h EHD,min is equivalent to the minimum lubricant volume conveyed, which occurred in the experiment at a speed of n = 200 rpm. The mean lubricant film thickness was calculated according to the approximation formula of H AMROCK / D OWSON for the line contact and is dependent on the normal force, see Figure 4-1. For the low normal force, the maximum possible smoothing was achieved as a result of the calculation result with a speed of n = 200 1/ min. For the two higher normal forces, the amount of smoothing increases, but the vertical line (minimum mean lubricant film thickness h EHD,min in the experiment) Figure 4-1: Prediction of surface roughness and change of surface roughness © WZL three normal forces. The hysteresis is most pronounced for the initial load at the lowest normal force and hardly recognizable at the highest normal force. In addition, the curve for a purely elastic contact calculation with the initial geometry is shown for comparison at the low normal force (Figure 4-2 left). The elastic-plastic contact calculation reduces the average lubrication pocket height, particularly with high solid contact ratios, which can be attributed to the smoothing of the roughness. To validate the contact calculation, the coefficients of friction from the experiment are finally compared with the results of the calculation, see Figure 4-2. The coefficient of friction is calculated on the basis of the solid contact ratio in the same way as in [MEVI19] with the same values for the solid and liquid coefficients of friction. The speed in the calculation results from the lubricant volume equilibrium and is equivalent to the mean lubrication pocket height. The elastic-plastic micro-contact calculation maps the renewed hysteresis between the speed run-down and run-up with each increase in the normal force. For the low normal force of FN = 1086 N, the hysteresis is overestimated by the calculation, resulting in a steeper drop in the friction coefficient. One reason for the deviation is the differences between the calculated and measured surface topography after the speed run, which are present after the deformation calculation. The differences in the surface topography can influence Science and Research 31 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 Figure 4-3: Friction Force Coefficients in Calculation and Experiment © WZL Figure 4-2: Calculation Results of Method © WZL [CZIC10] Czichos, H.; Habig, K.-H.: Tribologie-Handbuch. Tribometrie, Tribomaterialien, Tribotechnik. 3. Aufl. Wiesbaden: Vieweg+Teubner, 2010 [JACQ02] Jacq, C.; Nélias, D.; Lormand, G.; Girodin, D.: Development of a Three- Dimensional Semi-Analytical Elastic-Plastic Contact Code. In: J. Tribol., 124. Jg., 2002, Nr. 4, S. 653-667 [LEE12] Lee, Y.; Barkey, M.; Kang, H.: Metal Fatigue Analysis Handbook. Practical Problem-Solving Techniques for Computer-Aided Engineering. 1. Aufl. Waltham, MA: Butterworth-Heinemann, 2012 [LOHN16] Lohner, T.: Berechnung von TEHD Kontakten und Einlaufverhalten von Verzahnungen. Diss. TU München, 2016 [LÖPE15] Löpenhaus, C.: Untersuchung und Berechnung der Wälzfestigkeit im Scheiben- und Zahnflankenkontakt. Diss. RWTH Aachen University, 2015 [MEVI19] Mevissen, D.; Löpenhaus, C.; Bergs, T.: Calculation of mixed friction conditions in large-scale rolling-sliding contacts for different surface structures. In: Forsch Ingenieurwes, 83. Jg., 2019, Nr. 3, S. 351-366 [MEVI21] Mevissen, D.: Vorhersage der geometrischen Oberflächenveränderung im Wälzkontakt. Diss. RWTH Aachen University, 2021 [POPO10] Popov, V.: Kontaktmechanik und Reibung. Von der Nanotribologie bis zur Erdbebendynamik. Berlin: Springer, 2010 [PREX90] Prexler, F.: Einfluss der Wälzflächenrauheit auf die Grübchenbildung vergüteter Scheiben im EHD-Kontakt. Diss. TU München, 1990 [SAUE18] Sauer, B.; Magyar, B.; Oehler, M.: Schneckengetriebewirkungsgrade. Abschlussbericht zum Forschungsvorhaben FVA-Nr. 729 I, Heft 1226, Forschungsvereinigung Antriebstechnik e.V., Frankfurt a.M., 2018 [SCHU10] Schulz, J.; Holweger, W.: Wechselwirkung von Additiven mit Metalloberflächen. Renningen: Expert, 2010 [VOLG91] Volger, J.: Ermüdung der oberflächennahen Bauteilschicht unter Wälzbeanspruchung. Diss. RWTH Aachen University, 1991 Science and Research 32 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0004 the course of the solid contact ratio and thus the friction coefficient. The lower hysteresis with increasing normal force is correctly captured by the calculation. In addition, the friction coefficient curves intersect with decreasing speed, so that the calculation reflects the change from an increasing friction coefficient to a decreasing friction coefficient with increasing normal force. In summary, it can be stated that the running-in hysteresis of the friction coefficient can be predicted by the calculation method. In addition to the selection of the meshing parameters, the use of a stress-based yield criterion and the consideration of the lubricant pressure during the deformation of the half-space are key elements for the successful application. Acknowledgement The authors would like to thank the the Deutsche Forschungsgemeinschaft (DFG, German Research Foundation) [407625150 and 390969378] for providing the financial means to carry out the research project on which the results presented here are based. Literatur [AIRY62] Airy, G.: On the Strains in the Interior of Beams. In: Phil. Trans. R. Soc. London, 153. Jg., 1862, S. 49-79 [ASMI92] ASM International Handbook Committee: ASM Handbook. Friction, Lubrication, and Wear Technology. Bd. Nr. 18, Materials Park, OH: American Society for Metals, 1992 [BART10] Bartel, D.: Simulation von Tribosystemen. Grundlagen und Anwendungen. Wiesbaden: Vieweg+ Teubner, 2010 [BREC16] Brecher, C.; Löpenhaus, C.; Greschert, R.: Interaktion von fertigungs- und betriebsbedingten Grenzschichten imWälzkontakt. In: Tagungsband zur 57. Arbeitstagung “Zahnrad- und Getriebeuntersuchungen”. Aachen, 11./ 12.05.2016. Eigendruck des WZL-Getriebekreises, 2016 What does practice show? Even in trend analyses over several tens of thousands of operating hours, the oxidation value shows little change, although viscosity, neutralisation number, and/ or oxidation inhibitors indicate increased lubricant ageing (Example T1). Therefore, the significance of this standard, particularly for industrial lubricants, is rather limited. In determining oxidation in A/ cm at a specific point on the graph, a “band” at a wavenumber of 1710 cm -1 , the length change is calculated in cm compared to the base spectrum. In contrast, the oxidation index is based on an area calculation. The “oxidation index” is given as a dimensionless number and essentially corresponds to the increase in the oxidation area in cm 2 . The following examples demonstrate how further development of this method can lead to significantly improved results. After a brief introduction to the adjusted procedure for determining the oxidation index, the advantages of the method will be illustrated using typical examples. The first images show a typical example of a steam turbine, comparing the evaluation of the original FT-IR spectrum (fresh and used oil in transmission representation), as shown in Image 1. Science and Research 33 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 Oxidation Index - Determining the Ageing of Industrial Lubricants Carsten Heine* Presented at GfT Conference 2024 Every lubricant ages over its service life. For mineral oil-based lubricants, the oxidation is usually determined in accordance with DIN 51453. Alongside other analysis values, the oxidation value allows conclusions about the remaining service life of the oil or the progress of so-called oil ageing. The current version of DIN 51453 dates back to 2004, a time when neither today’s base oils nor current additive systems were available on the market. Furthermore, this method was developed for determining the ageing of engine oils. Due to a lack of alternatives, however, this method is still widely used for determining oxidation in almost all other applications, particularly for gas engine oils, hydraulic oils, and turbine oils. Keywords DIN 51453, mineral oil-based, engine oils, hydraulic, compressor, gear, oil changes Abstract * Carsten Heine OELCHECK GmbH Kerschelweg 28, 83098 Brannenburg, Germany Image 1: Representation of fresh and used oil in transmission dern base oil components) but rather analyse an area of the spectrum adjacent to the wavenumber 1710 towards the lower wavenumbers. As a result, the oxidation number of 3 A/ cm according to DIN is transformed into an oxidation index of 66 in this case. Many of the newer lubricants contain synthetic ester-based base oils. Not only do the anti-wear additives perform better in these oils, but they are also more temperature stable and age less. These synthetic components Science and Research 34 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 Since the original calculation of oxidation is based on the absorption representation, the next two images are also shown in this format. In Image 2, the determination of oxidation according to DIN 51453 is shown at the band of the wavenumber at 1710. Here, only the height of the peak in the difference spectrum is calculated and given as the result. In Image 3, it becomes clear that to determine the oxidation index, we do not rely on a single wavenumber (which could easily be influenced by additives or mo- Image 2: Representation of the difference spectrum in absorption (evaluation according to DIN 51453) Image 3: Representation of the difference spectrum in absorption (evaluation of the oxidation index) are advantageous, especially for oils in systems where oil changes are time-consuming and costly, or that operate at high temperatures. However, in oil analyses, these oils may cause difficulties because they oscillate around a wavenumber of 1740 and can thus influence the evaluation at a single wavenumber (1710 cm -1 ). Similarly, lubricants containing viscosity index improvers (VI-improvers) or special additives, which already show a significant peak area at 1710 cm -1 in the fresh oil, can influence the oxidation calculation. For modern lubricants, the oxidation value according to DIN 51453 is often inaccurate. For oils with ester components, determining oxidation according to DIN can be considerably challenging or even impossible. To provide a better indication of oil ageing for lubricants where DIN oxidation cannot be determined, viscosity increase, reduction in additive content, depletion of antioxidants, and the increase in AN (acid number), and in the case of engine oils, the decrease in BN (base number), have been considered. However, the new oxidation index provides a way to assess “oil ageing” more reliably and accurately via an oxidation value, which remains a focus for OEMs. The following section presents several typical examples from practice. The first example illustrates a case showing the significant consequences that can result if oxidation is not, or is unreliably, evaluated. Science and Research 35 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 Image 4: Example 1, Image 1 The FT-IR spectrum in Image 5 shows slight deviations in the region below (to the right of) the wavenumber 1710. These deviations are examined further. The FT-IR representation in absorption is used for calculation and visualisation. When oxidation is calculated according to DIN 51453, the result is 1.45 A/ cm, rounded by the software to 1. Using the new oxidation index method presented by OELCHECK, a result of 43 is obtained. As a dimensionless number, this requires interpretation. In the graph in Image 6, it is also evident that significant changes appear not only at the wavenumber 1710 but are more intense in the adjacent area, the basis of the new calculation method. Science and Research 36 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 This example concerns a steam turbine whose oil has been under our analytical supervision for an extended period. Images 4 and 5 show the current analysis, along with the trends from the last three analyses. In the data grid (Image 4), an increased copper content is noticeable, although it has remained stable across previous investigations. The sample from 2018 (second from the right) shows slightly notable values in the MPC test, foaming, and cleanliness class. Based on these values, temporary deep filtration and a partial replacement (about 10 % of the total volume) were performed, which returned the values to the previous trend level. It is notable that the oxidation value (according to DIN 51453) remains stable at a value of “1”. Image 6: Example 1, Image 3 Image 5: Example 1, Image 2 Example T1: Superheated steam turbine • DIN-Oxidation • 1.5 A/ cm • Oxidation-Index • 43 Example T1: Superheated steam turbine Example T1: Superheated steam turbine • DIN-Oxidation • 1.5 A/ cm • Oxidation-Index • 43 Example T1: Superheated steam turbine • DIN-Oxidation • 1.5 A/ cm • Oxidation-Index • 43 Example T1: Superheated steam turbine • DIN-Oxidation • 1.5 A/ cm • Oxidation-Index • 43 Example T1: Superheated steam turbine All FT-IR spectra of previous samples were recalculated using this method to examine changes over the investigation period. Image 7 clearly shows a surprising trend, indicating continuous oil changes over the entire monitored period. Example 2 In the second example (Images 8 and 9), the sample of an “initial analysis” of a gas turbine is shown. Other than the oil volume (approximately 20,000 litres), the only Science and Research 37 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 Image 7: Example 1, Image 4 Date Lab. No. Oxi DIN [A/ cm] Oxi-Index Ao ph [%] 06.02.2019 3926593 1 43 48.5 21.06.2018 3741784 1 35 56.4 27.12.2016 3274221 1 35 61.1 29.01.2016 3063875 1 27 64.9 24.02.2015 2833648 1 35 68.9 30.01.2013 2364102 1 34 67.6 31.01.2012 2141396 1 22 68.2 28.07.2010 1361871 1 18 81.5 Example T1: Superheated steam turbine Trend behaviour Date Lab. No. Oxi DIN [A/ cm] Oxi-Index Ao ph [%] 06.02.2019 3926593 1 43 48.5 21.06.2018 3741784 1 35 56.4 27.12.2016 3274221 1 35 61.1 29.01.2016 3063875 1 27 64.9 24.02.2015 2833648 1 35 68.9 30.01.2013 2364102 1 34 67.6 31.01.2012 2141396 1 22 68.2 28.07.2010 1361871 1 18 81.5 Example T1: Superheated steam turbine Trend behaviour Image 8: Example 2, Image 1 Example T2: Gasturbine Oi l: Turbine Oil ISO VG 46, 20.000 Litres Example T2: Gasturbine Oi l: Turbine Oil ISO VG 46, 20.000 Litres Date Lab. No. Oxi DIN [A/ cm] Oxi-Index Ao ph [%] 06.02.2019 3926593 1 43 48.5 21.06.2018 3741784 1 35 56.4 27.12.2016 3274221 1 35 61.1 29.01.2016 3063875 1 27 64.9 24.02.2015 2833648 1 35 68.9 30.01.2013 2364102 1 34 67.6 31.01.2012 2141396 1 22 68.2 28.07.2010 1361871 1 18 81.5 Example T1: Superheated steam turbine Trend behaviour Example T2: Gasturbine Oi l: Turbine Oil ISO VG 46, 20.000 Litres Example 3 In the third example, samples are shown from a steam turbine that, unlike the first two examples, is operated with a slightly EP-additive turbine oil. This case also serves to check whether the oxidation index provides similarly good results for more heavily formulated oils. The investigated sample comes from a steam turbine where there was a suspicion that the oil, which had been in use for around 30,000 hours, might already be showing signs of oxidation. The slightly more extensive standard analysis (Image 10) already provided initial indications of oil changes, with a significantly worsened water separation capability, deteriorating foam behaviour, and a reduction in amine antioxidants in the RULER test. The oxidation value (according to DIN 51453) was also somewhat elevated, with a value of “3”. However, with such small values, interpretation can be challenging. The FT-IR spectrum (Image 11) also clearly shows changes in the area to the right of the wavenumber 1710 A/ cm. In Image 12, it can be seen that, in this case, the wavenumber 1710 A/ cm lies directly at the centre of the IR changes, already indicating that it is pointing in the right direction. Science and Research 38 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 information available was that the turbine oil was ISO VG 46. No details regarding the oil service life or the total operating time of the system were provided. The oxidation value of “2” without any reference to oil service life is initially not very informative. However, the FT-IR spectrum (Image 9) reveals not only slight changes in the area around the wavenumber 1710 A/ cm, but also significant changes compared to fresh oil in relation to the phenolic antioxidants (wavenumber range at around 3,600 A/ cm). The RULER test was not requested for this sample, so no statement could be made in this regard. However, the calculation of the degradation of phenolic antioxidants via the FT-IR spectrum was very clear. Only about 16 % of the phenolic AO content remained when compared to the fresh oil. Additionally, the oxidation index calculated for this sample also produced a significant result. The oxidation index had risen to a value of 43. The recommendation in this case was to also examine the deposit formation tendency via the MPC test and to derive further measures from this if necessary. During the tank inspection, noticeable deposits were found on the tank walls at the usual filling level. Due to the unclear data situation and a scheduled shutdown, no further investigations were commissioned, and an oil change, along with a complete system cleaning, was carried out after another two months. Image 9: Example 2, Image 2 Example T2: Gasturbine Oi l: Turbine Oil ISO VG 46, 20.000 Litres Oxidation DIN [A/ cm] Oxidation-Index Ao ph [%] 2 43 16.3 (No Trend existing) Example T2: Gasturbine Oi l: Turbine Oil ISO VG 46, 20.000 Litres Oxidation DIN [A/ cm] Oxidation-Index Ao ph [%] 2 43 16.3 (No Trend existing) Example T2: Gasturbine Oi l: Turbine Oil ISO VG 46, 20.000 Litres Oxidation DIN [A/ cm] Oxidation-Index Ao ph [%] 2 43 16.3 (No Trend existing) Example T2: Gasturbine Oi l: Turbine Oil ISO VG 46, 20.000 Litres Oxidation DIN [A/ cm] Oxidation-Index Ao ph [%] 2 43 16.3 (No Trend existing) Science and Research 39 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 Image 10: Example 3, Image 1 Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Control hydraulics, EP turbine oil ISO VG 46, 600 litres SST 400 Sample and Cap Infrared spectrum Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Control hydraulics, EP turbine oil ISO VG 46, 600 litres SST 400 Sample and Cap Infrared spectrum Image 11: Example 3, Image 2 Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Control hydraulics, EP turbine oil ISO VG 46, 600 litres Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Control hydraulics, EP turbine oil ISO VG 46, 600 litres Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Control hydraulics, EP turbine oil ISO VG 46, 600 litres Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Control hydraulics, EP turbine oil ISO VG 46, 600 litres SST 400 Sample and Cap Infrared spectrum The value determined by the oxidation index, “66”, clearly shows that substantial oil ageing has already occurred. Science and Research 40 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 However, Image 13 also shows that there are significant changes in the area adjacent to the wavenumber 1710 A/ cm. Image 12: Example 3, Image 3 Oxidation DIN = 3 A/ cm Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Oxidation DIN = 3 A/ cm Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Image 13: Example 3, Image 4 Oxidation DIN = 3 A/ cm Oxidation index = 66 Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Oxidation DIN = 3 A/ cm Oxidation index = 66 Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Oxidation DIN = 3 A/ cm Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Oxidation DIN = 3 A/ cm Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Oxidation DIN = 3 A/ cm Oxidation index = 66 Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Oxidation DIN = 3 A/ cm Oxidation index = 66 Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Oxidation DIN = 3 A/ cm Oxidation index = 66 Example: Steam Turbine with heavily oxidised oil Siemens SST 400, Copntrol hydraulics, EP additivated oil, ISO VG 46, 600 litres Summary More than 15,000 samples of turbine oils, hydraulic oils, compressor oils, and gear oils have now been investigated and evaluated using this method. The experiences gathered from these samples clearly show that oxidative changes can be detected and assessed at a much earlier stage. The oxidation index is primarily an early indicator of oxidative changes in industrial oils. The large number of samples has enabled us to establish initial limit values for different applications. These are currently being continuously reviewed, refined, and adjusted. This alternative method of evaluating the FT-IR spectrum regarding oxidation provides meaningful values for assessing oxidation in modern lubricants. By offering more understandable and stable trend values, this method significantly improves operational safety for plant operators. Through trend progression, a much better prediction of the remaining oil life in relation to oxidation can be made. In conjunction with the classical values of oil analysis, further meaningful tests (e.g., MPC, RULER) can be recommended, if necessary, to assess oil changes accurately and identify causes, as well as find solutions. This method has also attracted interest in the DIN working group for used lubricants. Currently, a task force from this working group is evaluating the incorporation of this method into a DIN standard. References: DIN 51453: 2004-10 Science and Research 41 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 DOI 10.24053/ TuS-2025-0005 News 42 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 The 66 th Tribology Conference, organized by the German Society for Tribology (GfT) will take place for the first time in Wernigerode, a beautiful city in the Harz Mountains. It will be held from September 29 to October 1 at the HKK Hotel, which has turned out to be perfectly suited for this conference. The conference program will include more than 80 scientific lectures in six parallel sessions on cutting-edge tribological topics. It will also include a poster session and a trade exhibition in the foyer of the conference hotel. As in previous years, there will be a special session with presentations from the DFG priority program “Fluidfree lubrication systems with high mechanical loads”. The topics revolve around e.g. dry lubrication of gears, rolling contacts, worm gears and screw machines. In another special session lectures from the research field “Tribology” will be given with subjects from the areas of wear reducing coatings, oil lubrication, additives and tribological measures to increase efficiency. The research field is funded as part of the energy research program of the Federal Ministry for Economic Affairs and Climate Action and thus contributing to the energy transition in Germany. Key Topic: Cost reduction by tribology Tribological measures to reduce CO 2 emissions and increase sustainability are often associated with energy cost reduction potentials. Furthermore, tribological Gesellschaft für Tribologie 66 th German Tribology Conference for the first time in Wernigerode Wernigerode, timbered houses and castle Market Square and Town Hall Plenary Session Poster Session Sessions from funded programs give scientists the opportunity for personal discussions beyond the boundaries of the individual program and strengthen the character of the conference as an important working meeting of leading experts in the field of tribology. News 43 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 measures can also achieve additional cost reductions in the application. The lectures of this session are intended to show the savings potential that result from the consistent application of tribological know-how in the design, development, manufacturing and application of goods. Special Topic: Interaction of additives with metal surfaces Lectures on lubricants and lubrication technology as well as on machining and forming technology will be a major part of conference program. In both areas, the interaction of additives with metal surfaces is of crucial importance. This will be taken into account by a separate session on this topic. Contributions on reaction layers, the acting mechanisms, as well as the corresponding theories and models are expected. 4 th TriboSlam Tribology is boring? A scientific conference can’t be fun? Not at all! Every year since 2022, the German Tribology Conference has proven how much fun tribology is and how exciting research can be. Let yourself be inspired by the slams of the last few years on GfT’s You- Tube channel https: / / www.youtube.com/ @GfTeV. If you have an idea, we cordially invite you to participate in the 4 th TriboSlam! Honors and Awards Every year, the Georg Vogelpohl Medal of Honor is awarded to an outstanding personality in the field of tribology. The award ceremony will take place during the plenary session. Also, in the plenary session young scientists and engineers who have achieved above-average performance in their bachelor, master and doctor thesis will be awarded by GfT Sponsorship Awards. TriboSlam Every year, the GfT awards the prize “Tribology is everywhere”, donated by the company Werner Stehr Tribologie. The prize is intended to sharpen the view on tribological phenomena in everyday life and can be awarded for an observation captured in a picture, a newspaper article, but also for a scientific paper. Usually, the submitted topic is presented by the winner in a plenary lecture during the closing event in a vivid way, which is always a highlight of the conference. The GfT is looking forward to welcoming many participants to the new conference location. Abstracts for oral presentations and posters can still be submitted via the GfT website. Further information can be found in the call for papers and on https: / / www.gft-ev.de/ wp-content/ uploads/ 2025-Call-for-Papers.pdf News 44 Tribologie + Schmierungstechnik · volume 72 · issue 1/ 2025 The European Conference on Tribology - ECOTRIB - is a biennial conference organized in collaboration between the Austrian, Italian, Slovenian and Swiss tribology societies. Previous Ecotrib meetings have taken place in Ljubljana (2007, 2017), Pisa (2009), Vienna (2011, 2019), Turin (2013), Lugano (2015), and Bari (2023). The 9 th Ecotrib meeting will organized by Swiss Tribology in Zurich, Switzerland, in the main building of ETH Zurich. Topics to be covered in the meeting include • modeling, simulations, and machine learning in tribology • nano and microtribology • tribochemistry • tribocorrosion Please check for further information: https: / / www.ecotrib25.ch/ Swiss Tribology ECOTRIB 25 ETH Zurich, Zurich, Switzerland, 28-30 JULY 2025 • biotribology and biotribomaterials • coatings and surface engineering • friction, wear and contact mechanisms • green tribology • industrial, automotive and manufacturing case studies • lubrication and lubricants, including solid lubricants Checklist Author information Corresponding author: F Mailing address F Telephone and fax number F eMail All authors: F Academic titles F Full name F ORCID (optional) F Research instititute / company F Location and zip code Length F Approximately: 3,500 words Data F Word and pdf documents (both with images + captions) F Additionally, please send images as tif or jpg / 300 dpi / Please send vector data as eps Manuscript F Short and concise title F Keywords: 6-8 terms F Abstract (100 words) F Numbered pictures/ diagrams/ tables (please refer to the numbers in your text) F List of works cited After the typesetting is completed, you will receive the proofs, which you are requested to review and then give your approval to start the printing process. Changes to the manuscript are no longer possible at this stage. Please also consider The editors and the publisher assume that the authors are authorized to publish all data used, that the provided texts and all visual material (images/ pictures/ illustrations) do not violate any (copy)rights of a third party, and that, where necessary, source references are provided for visual material. In cases of doubt, please obtain a printing permission from the copyright holder. Editors and publisher cannot assume liability for potential copyright infringements. Open Access Free access to knowledge is important to us. That is why you also have the opportunity to make your contribution immediately available digitally to all interested parties. This not only benefits you with an increased reach, but also researchers worldwide. In order to guarantee the high quality and substantial indexing, we are unfortunately unable to offer this service free of charge. 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ISSN 0724-3472 Science and Research www.expertverlag.de Felix Schlegel, Marius Hofmeister, Dino Osmanovic, Katharina Schmitz Swelling, wear and property changes of high-performance polymers in oil-hydraulic tribological contacts Jan Euler, Georg Jacobs, Timm Jakobs, Thomas Decker, Noah Smeets, Julian Röder Influence of wear and manufacturing inaccuracies on the performance of a conical plain bearing main bearings for wind turbines Dieter Mevissen, Sebastian Sklenak, Christian Brecher, Thomas Bergs Elastic-plastic micro contact calculation for large scale, lubricated rolling-sliding contacts Carsten Heine Oxidation Index - Determining the Ageing of Industrial Lubricants