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Tribologie und Schmierungstechnik EDITOR IN CHIEF MANFRED JUNGK 1 _ 24 VOLUME 71 eOnly SPECIAL ISSUE Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie An Official Journal of Österreichische Tribologische Gesellschaft An Official Journal of Swiss Tribology eOnly Special Issue 1 | 2024 Volume 71 Editor in chief: Dr. Manfred Jungk Tel.: +49 (0)6722 500836 eMail: manfred.jungk@mj-tribology.com 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 978-3-381-12442-8 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 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0022 The 24 th International Colloquium Tribology took place from January 23 rd to 25 th 2024 at Technische Akademie Esslingen and as in the past 23 events presented an ideal communication platform for representatives from industry and science. Global challenges regarding energy supply, resource conversation and the urgent need to take completely new pathways in transport and production to combat global warming are eminent in today’s topics. Tribology, as the science of friction, wear and lubrication, is uniquely positioned to address these challenges due to its interdisciplinary nature. The program was built around 5 topics: New trends in lubricants and additives, Coatings, surface interactions and underlying mechanisms, Machine elements and their applications in tribology, Computational methods and digital transformation in tribology, Test and measurement methodologies. The steering committee consisting of Nicole Dörr, Katharina Völkel, Max Marian & Carsten Gachot invited excellent speakers for keynote and plenary topics on mobility, data handling and efficiency enhancement of machine elements. In this e-only issue you will find exemplarily in detail a publication presented at the 24 th International Colloquium Tribology with the title “The Effects of Applying the Tribological Compound TZ NIOD - reversing wear”. TZ NIOD is a complex mixture of silicate material powder with particle sizes ranging from 5 to 50 micrometres. The basis is made up of finely distributed and divided particles of Serpentinite. Wind Tribology was often a topic at the events of International Colloquium Tribology back to the days where micro pitting was a big problem at the start of the millennium. Micro pitting has been long solved, but other topics surfaced during the upscaling from 50 KW turbines to now reaching double digit MW’s. Therefore, you will find three recent articles on wind turbine tribology as the focus theme. Jörg Loos et.al. report on “Friction induced WEC-formation at high loads”. White Etching Cracks occur also in wind turbine rolling element bearings. The article describes a frictional WEC lifetime model, which was derived from a large number of previous bearing tests, providing an explanation for different failure behaviours in the WEC-tests. Ivan Grozev et.al. reports on “Simplified tribological approach for predesign of wind turbine bearing cases, combined with model test investigation”. They learned that 2-Disc tribometer may be used to simulate 100 % sliding, making comparison of different slippage rate easier by only using one tribometer, thus reducing measurement inaccuracies caused by different type of machines (e.g., pin on disc) or operator’s influence Jan Euler et. al. report on “Feasibility study on preloaded flexible conical plain bearings as wind turbine main bearings”. Please enjoy reading this e-only issue and remember Tribology is everywhere. Your editor-in-chief Manfred Jungk 24 th International Colloquium Tribology and Focus Wind Turbine Events 2 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 Events We look forward to your contribution! The scientific journal Tribologie und Schmierungstechnik (TuS) is one of the leading publications for tribological research in Germany, Austria and Switzerland. As the official journal of the Society for Tribology (GfT) in Germany, the Austrian Tribological Society (ÖTG) and Swiss Tribology, the issues provide information on research from industry and science, current events and developments in the specialist community. Further information on the journal and publication: https: / / elibrary.narr.digital/ xibrary/ start.xav? zeitschriftid=tus&lang=en Date Place Event ► 26.04. - 29.04.25 Copenhagen, Denmark ELGI 35 th Annual General Meeting ► 13.05. - 15.05.25 Brannenburg, Germany Oildoc Conference ► 18.05. - 22.05.25 Atlanta, Georgia (USA) 79 th STLE Annual Meeting & Exhibition ► 28.07. - 30.07.25 Zürich, Switzerland European Conference on Tribology - ECOTRIB ► 29.09. - 01.10.25 Wernigerode, Germany 66. GfT Conference Tribology Contents 3 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 Tribologie und Schmierungstechnik Tribology - Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie An Official Journal of Österreichische Tribologische Gesellschaft An Official Journal of Swiss Tribology Volume 71, eOnly special issue 1/ 2024 January 2025 4 Philipp Harrer, Dmitrii Svetov, Patrick Eisner, Maximilian Lackner The Effects of Applying the Tribological Compound TZ NIOD - reversing wear 1 Editorial 24 th International Colloquium Tribology and Focus Wind Turbine 2 Events TAE-Colloquium Tribology 2024 10 Jörg Loos, Wolfram Kruhöffer, Daniel Merk, Toni Blaß, Jörg Franke Reibungsbedingte WEC-Bildung bei hohen Lasten 19 Ivan Grozev, Sagar Dalal, Nazlim Bagcivan, Serhan Bastuerk, Christian Lueffe, Thomas Stahl Simplified tribological approach for predesign of wind turbine bearing cases, combined with model test investigation 26 Jan Euler, Georg Jacobs, Timm Jakobs, Julian Röder Feasibility study on preloaded flexible conical plain bearings as wind turbine main bearings Wind-Tribology 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). TAE-Colloquium Tribology 2024 4 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0023 The Effects of Applying the Tribological Compound TZ NIOD - reversing wear Philipp Harrer, Dmitrii Svetov, Patrick Eisner, Maximilian Lackner* Presented at 24 th International Colloquium Tribology (TAE) | submitted: 14.04.2024 accepted: 6.05.2024 TZ NIOD is a complex mixture of silicate material powder, to mix with lubricant and apply to moving parts of a device under operation to improve its tribological properties. It reaches worn surfaces and highly loaded friction points to react with the material creating a modified surface layer. Empirical analyses with TZ NIOD applied to used piston air compressors required a down time of only 60 minutes and resulted in reducing the average power consumption per pressure vessel filling cycle by 20.7 W (-7.8 %) and reducing the average filling time of the pressure vessel by 3.9 seconds (-5.1 %). Keywords Tribology, green tribology, sustainability, Tribo-System, Nanoparticles, Friction, Wear, TZ NIOD Abstract * Ing. Philipp Harrer, Msc 1 Dmitrii Svetov 2 Dipl.-Ing. Dr. mont. Patrick Eisner, MSc 1 PD Di Dr. techn Maximilian Lackner, MBA 1 (corresponding Author) 1 UAS Technikum Wien, Industrial Engineering, 1200 Vienna, Austria 2 Dmitrii Svetov www.tribo.at, 1150 Vienna, Austria 1 Introduction Tribology’s economic and technical relevance in terms of energy loss, material deterioration and waste has long been accepted and has recently been augmented with sustainable viewpoints such as environmental awareness, longer life of device, reducing waste and enhancing the quality of life. Due to the vast potential and relevance in various sectors, improvements in the field of tribology remain of high importance. [1] Wear due to friction is one of the leading causes of damage to components and thus the leading cause for the failure of equipment and devices. [2] Friction cannot be avoided when components in motion are in direct contact, it can however be reduced. Therefore, measures must be taken to reduce friction and thus the amount of wear and the multitude of the negative side effects of it. It not only causes wear and initiates the need for replacement of parts but friction has much more extensive effects on humanity. Friction has been identified as a potential contributor to global warming due to the loss of energy to overcome friction. One fifth of the global energy consumption is required to overcome friction alone. Reducing friction could have a significant impact on reducing the global CO2 emissions. Study shows that the global transportation sector consumes 200 billion liters of fuel only to overcome friction. Roughly one third of the energy consumption of the transportation sector is required to overcome friction. New technology for reducing friction is allegedly capable of cutting losses due to friction in half which, from a global perspective, could potentially eliminate 960 million tons of CO 2 emissions. [3] Reducing friction not only reduces the energy consumption of devices but also reduces wear as a whole, which improves the useful life of components or systems and reduces the generation of heat, which would allow a downsizing of the component itself. This would result in longer use periods of devices at a much lower resource and energy consumption. [4] Thus, the sector of tribology has enormous potential in reducing the global energy consumption, depletion of resources as well as generation of greenhouse gases. The results of research and development of novel methods of reducing friction must be taken into consideration and can contribute significantly to a sustainable future. One potential contributor to reduce friction is TZ NIOD [5], which is a tribological compound which shall be applied to moving parts with the goal of reducing friction, energy consumption, renewing worn out surfaces, increasing the service life of the entire device, reducing the temperature, reducing the coefficient of friction as well as reducing the rate of wear. It consists of a complex mixture of silicate material powder, specifically serpentinite, which uses oil or grease as a transport medium to reach worn surfaces and highly loaded friction points. These particles allegedly react with the material under the influence of temperature and pressure creating a modified surface layer. The aim is to analyse the effects of TZ NIOD on tribosystems and devices it is applied to. Empirical analyses were performed in alignment with the tribological test chain, specifically machinery tests, analysing the effects of an application of TZ NIOD on piston compressors. These analyses consisted of temperature and power consumption measurements during various modes of operation of the compressor and different modes of applying TZ NIOD to the compressor. 2 Tribological Compound - TZ NIOD TZ NIOD [5] is a novel agent to improve properties of friction partners. TZ NIOD is a complex mixture of silicate material powder with particle sizes ranging from 5 to 50 micrometers. The basis of TZ NIOD is made up of finely distributed and divided particles of Serpentinite. It consists of nanoparticles which must be dispersed in oil or grease and the intensity of its penetration into the material surface is proportional to the pressure and temperature of contact zones. It is important to state that TZ NIOD is neither a modifier, an additive for lubricants nor a lubricant on its own. It is a tribological compound intended to be combined with a lubricant and applied for only a limited amount of time. The lubricant such as grease or oil acts as a transport medium to transport the TZ NIOD particles towards the highly loaded friction points of tribological systems such as gears, motors, compressors, and the components thereof. One of the most advantegous claim of TZ NIOD is its alleged capability of renewing mechanisms subjected to friction and wear while the mechanism is in operation. This means the tribological compound TZ NIOD can allegedly be applied to a device in operation without the need for disassembling components and without any significant down-time of the device. An application of TZ NIOD on tribological systems apparently improves the friction points by reducing the coefficient of friction as well as the rate of wear. It allegedly improves the properties of the interacting surfaces and is embedded in the material structure through which its effect unfolds after an application of TZ NIOD, even once TZ NIOD has been removed from the tribo-system 2.1 Claimed Benefits of TZ NIOD Svetov [2] claims that the nanoparticles of TZ NIOD accumulate in worn areas of contact zones of tribo-systems due to the higher surfaces roughness. Thus, the oil and grease act as a transporting agent of the TZ NIOD particles to the areas of highest wear. The increased friction, temperature and higher pressure, due to wear, stimulate the penetration of TZ NIOD particles into the contact surface resulting in a mending and self-healing effect. This self-healing effect results in a restoration of the friction partners of the tribo-system and shall have the following positive effects. • reduce the coefficient of friction • reduce energy losses due to friction • lower temperature increases due to friction • higher resistance against wear • ability to operate tribo-systems without lubrication for short periods of time 2.2 Working Principles of TZ NIOD An application of TZ NIOD is performed directly on the device during its operation, within the devices operating boundaries, and consist of three phases [6]. Phase 1: activation of TZ NIOD and the surfaces in the friction zone Phase 2: “diffusion” of TZ NIOD into the surface layers of the metal Phase 3: “diffusion” of TZ NIOD from the surface layer deep into the metal. In phase one, finely dispersed TZ NIOD particles are transported to the areas of wear via the oil and grease it is dispersed in. The friction in the contact zone grinds down the TZ NIOD particles and forms activated TZ NIOD particles. The particles have an abrasive effect which polishes the contact areas due to the hardness of the silicate material. These abrasive particles remove oxide layers from the metal and react with it under the influence of temperature and pressure though which it diffuses into the metal structure of the tribo-system. Due to the diminishing size of the particles the effect of TZ NIOD gradually depreciates and the gradually shrinking TZ NIOD particles turn into ultra-fine abrasive particles. These particles are transported through the device until the particles are too small to have any additional benefit. Therefore, the presence of TZ NIOD in the device is only relevant for a certain period of time. The second phase of the process begins when a sufficient concentration of activated TZ NIOD particles is formed in the contact zone. The website of the project sponsor claims that TZ NIOD particles diffuse into the metal structure and are embedded. This results in a modified surface layer with increased hardness and higher resistance against wear. The surface now consists of a compound with different structures. The process of phase 2 continues until the entire surface of the contact area is saturated with TZ NIOD particle. Since the metal structure is saturated with TZ NIOD at the end of Phase 2 the lubricant containing the remaining finely dispersed and grinded TZ NIOD can be removed from the tribological system and replaced by a fresh lubricant. According to Yu et al. it has been observed that fine particles of serpentine, which is the main component of TZ NIOD, mixed with lubricants form very hard and super-lubricious oxide layers on worn metallic surfaces which are TAE-Colloquium Tribology 2024 5 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0023 3) Applying TZ NIOD to the device in the application phase. 4) Replacing the TZ NIOD-Oil mixture with fresh lubricant and operating the device for an extended period of time in its standard operating mode, at a duty cycle of 60 %, in the “Running-In” phase. 3.1 Applying TZ NIOD to the Piston Air Compressor The application of TZ NIOD must be designed specifically for each device, because each type of device possesses different friction areas. The advantage of TZ NIOD is that it can be applied while the device is in operation and does not require a disassembly of the components or tribo-systems. To design the application procedure, the areas of highest friction must be defined. The areas of highest friction of the piston air compressor are the crankshaft, the bearings of the connecting rod, the piston rings as well as the valves. The application method must allow a transport of TZ NIOD towards the defined areas. The listed friction points can be grouped into two parts of the compressor, the cylinder head comprising the valves, the piston and piston rings as well as the crankcase and the components therein. To treat of the components of the crankcase, the TZ NIOD-Oil mixture is filled into the oil pan instead of the normal compressor oil. It is important to note that the TAE-Colloquium Tribology 2024 6 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0023 capable of lowering friction and wear. In addition, the formation of diamond-like carbon (DLC) layers has been observed on surfaces subjected and exposed to wear, for lubricants containing serpentine which resulted in outstanding tribological properties of the tribo-system. Both sources endorse the effect claimed of TZ NIOD. The third phase continues after the removal of the lubricant containing TZ NIOD. The device must continue to operate and the TZ NIOD particles embedded in the metal structure continue to “diffuse” into deeper levels of the metallic structure. The duration of the third stage can be several hundred hours of normal operation of the device without the need of applying additional TZ NIOD. Such effects of creating relatively thick, hard and wear resistant oxide layers have been observed by Yuansheng et al. with serpentine containing nanoparticles dissolved in lubricants [7]. Yuansheng et al. states that a mechanochemical oxidation and reaction of iron oxides was responsible for the formation of the wear resistant layer. A thickness of the protective layer of 8 to 10 µm was measured. An application of TZ NIOD on a device consists of the following stages: 1) The oil in the oil pan of the device must be removed and rinsed with fresh oil in case the oil pan is contaminated with residue from the old oil. 2) The TZ NIOD powder is mixed with fresh oil in the correct ratio and filled into the oil pan of the device. 3) The application phase is started and shall be carried out by operating the device. Phases 1 and 2 of the above-described working principles of TZ NIOD occur in the application phase. The operating conditions are used to apply the load. The load shall be 70 % of the maximum load and shall be applied for a range of 30 minutes up to 50 hours. If the device is intended to move in multiple directions, then the device shall be operated in all directions during the application phase. The duration of the application phase depends on the size and load cases of the device as well as on the speed of rotation. Slowly rotating devices require a longer application phase. 3 Emperical Analysis To test the effect of TZ NIOD on real equipment in a machinery test, TZ NIOD was applied to a more than 40year-old used piston air compressor. The application of TZ NIOD on the piston air compressor consisted of the following four stages: 1) Removing the initial oil from the device. 2) Filling the TZ NIOD - Oil mixture in the proper ratio into the oil pan of the device. Table 1: General Data of the Compressor Figure 1: Piston Air Compressor of the Empirical Analysis ! " operation of the compressor is started immediately after introducing the TZ NIOD-Oil mixture into the oil pan. This is of particularly high importance to prvent TZ NIOD particles from sinking to the bottom of the oil pan and losing their effectiveness. The cylinder head and the components therein of such piston air compressors are typically not treated with oil. This part of the compressor does not have an oil sump or other methods of lubrication. Therefore, a method of applying the mixture of oil and TZ NIOD was designed. The applicating of the mixture of oil and TZ NIOD was performed via the intake air of the compressor by removing the air filter and trickling the TZ NIOD-Oil mixture into the air inlet over a prolonged period of time. The compressor must be operated during the application process in order for the mixture of oil and TZ NIOD to be sucked into the cylinder head. Additionally, the pressure line, the connection to the pressure vessel, must be removed to prevent filling the pressure vessel with the mixture. The application of the TZ NIOD - Oil mixture was conducted in the application phase represented in the temperature over time diagram of Figure 2. As stated, the TZ NIOD - Oil mixture was filled into the oil pan of the device and the air filter as well as the pressure line of the piston air compressor were disconnected prior to starting the application phase. The application phase began with continuous operation of the device for 20 minutes (section 1 of Figure 2) in which the TZ NIOD - Oil mixture was applied to the air intake of the device exposing the cylinder head and the components thereof to TZ NIOD. Next, for the second phase (section 2 of Figure 2) the device was operated continuously under full load for 40 minutes with the air filter mounted and the pressure line connected to the pressure vessel. The third phase consisted of 3 hours and 20 minutes of the device’s standard discontinuous operation at a utilization rate of 60 % (section 3 of Figure 2). The duration for section 3 (discontinuous operation) was selected to result in a total duration of the application phase of 4 hours. Finally, the TZ NIOD- Oil mixture was replaced by fresh lubricant (section 4 of Figure 2) and the running-in phase of TZ NIOD was initiated. During the running-in phase the device was operated in its standard operating mode. The reason for the running-in phase is to initiate phase 3 of the working principles of TZ NIOD, and thus for the “diffusion” of TZ NIOD from the surface layer deep into the metal to occurs. For the empirical analysis the running-in phase consisted of operating the piston air compressor for 100 hours in its standard discontinuous operating mode at a utilization rate of roughly 60 %. The utilization rate was achieved by filling the pressure tank to its maximum pressure level of 1 MPa for 3 minutes followed by 2 minutes of resting to empty the pressure tank. As illustrated in Figure 2, The application of TZ NIOD resulted in a down time of 60 minutes (sum of section 1 and 2) and only required the air filter as well as the pressure line of the compressor the be removed. No additional disassembling of the compressor or modifying its components or tribo-systems was required for the application of TZ NIOD. 3.2 Results of TZ NIOD applied to the Piston Air Compressor The pressure of the pressure vessel and filling time, the power consumption of the motor, as well as the temperature of the cylinder head were recorded in the initial state (before applying TZ NIOD), directly after the application phase as well as after the running-in phase with a duration of 100 hours. TAE-Colloquium Tribology 2024 7 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0023 Figure 2: description of the sections of the application phase TAE-Colloquium Tribology 2024 8 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0023 Figure 3: Pressure-Time Diagram of initial, after application and after running-in Figure 4: Detail of Power Consumption filling the Pressure Vessel after running-in with TZ NIOD ! "# $% & % $ Table 2: Result of Filling the Pressure Vessel after Running-In phase with TZ NIOD During the empirical analysis it was observed that the pressure over time characteristic of the compressor improved noticeably, see Figure 3. A clear difference can be noticed between the characteristic after running-in (blue line in Figure 3) compared to the initial state (black line in Figure 3) and the state directly after applying TZ NIOD (green line in Figure 3) Further, the application of TZ NIOD positively affected the filling time and the power consumption, as compared in Table 2. The average power consumption of each filling cycle during a typical discontinuous operating mode for filling the pressure vessel was reduced by 20.7 W, which represents a reduction of 7.8 percent. The average time to fill the pressure vessel was shortened by 3.9 seconds, which represents a reduction of 5.1 percent. The power consumption characteristic while filling the pressure vessel is illustrated in Figure 4. The power consumption after running-in is represented by the blue graphs in Figure 4. The blue graphs are noticeably lower than the graphs of the initial state, represented in black. 4 Conclusion The results of the empirical analyses conclude that TZ NIOD is capable of unfolding its positive effects when applied to devices in operation and within the specified operating condition of the device. An application of TZ NIOD must be tailored to the specific device. The application of TZ NIOD on the piston air compressor resulted in a total down time of only 60 minutes. Overall, it can be concluded that the positive effects of TZ NIOD, on the device it is applied to, comprises of lowering the power consumption by 7.8 %, increasing the efficiency by lowering the filling time of the pressure vessel by 5.1 %. This indicates that the worn-out surfaces of the device were regenerated which contributed to decreasing the temperature in operation and increasing the devices service life. Due to the observed positive and confirmed effects of TZ NIOD on devices it was applied on, leading to a decrease of power consumption, lower energy demand, faster filling times and thus a decrease of temperature due to the shorter operating times, it can be underlined that an application of TZ NIOD is of high potential. The application of TZ NIOD is capable of significantly contributing to the field of green tribology and sustainability by reducing costs as well as potentially contributing to energy savings, material savings, reducing waste, and a longer life of devices. Its application is very simple and can be performed on a multitude of devices without the need for significant down times. This is due to the fact that it can be applied to the device in operation. References [1] J. P. Davim, Progress in Green Tribology: Green and Conventional Techniques. Berlin: De Gruyter Oldenbourg, 2017. [2] Popov, Valentin. L. (2015): Kontaktmechanik und Reibung : Von der Nanotribologie bis zur Erdbebendynamik. 3 rd edn., Berlin Heidelberg: Springer Publishers. [3] Holmberg K, Andersson P, Erdemir A, Global energy consumption due to friction in passenger cars. Tribology International 47 (2012) 221-234 [4] Wedeven, Vern (2022): Introducing friction, wear, and lubrication in the revolutionary tribology, AZoM.com. [online]https: / / www.azom.com/ article.aspx? ArticleID=21785 [April 5, 2023] [5] D. Svetov, “Tribologie in Österreich,” Tribo.at, http: / / tribo.at/ (acc. Apr. 18, 2023). [6] D. Svetov, “Die tribotechnische Zusammensetzung von Niod - Prozesse,” Tribo.at, http: / / www.tribo.at/ prozes.html (accessed Apr. 18, 2023). [7] Yuansheng, J./ Shenghua, L./ Zhengye, Z./ He, Y./ Feng, W. (2004): In situ mechanochemical reconditioning of worn ferrous surfaces. In: Tribology International. vol. 37. P. 561-567. TAE-Colloquium Tribology 2024 9 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0023 Wind-Tribology 10 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 Einleitung Bei White Etching Cracks (WECs), auch bekannt als White Structure Flaking oder Brittle Flaking, handelt es sich um einen Ermüdungsmechanismus, der sich nicht mit der klassischen Wälzlager-Lebensdauertheorie berechnen lässt. Dabei kommt es unter der Oberfläche zu Rissnetzwerken an und in weiß anätzenden Phasen [Hol15, Eva16]. Betroffen können sowohl durchals auch einsatzgehärtete Lager [Bla17] und alle Wälzlagertypen [Loo17] unabhängig von der Art der Schmierung (Öl oder Fett) [Loo16a] sein. Die weiß anätzenden DOI 10.24053/ TuS-2024-0024 Reibungsbedingte WEC-Bildung bei hohen Lasten Jörg Loos, Wolfram Kruhöffer, Daniel Merk, Toni Blaß, Jörg Franke* Wälzlager können in seltenen Fällen deutlich vor der berechneten Lebensdauer mit White Etching Cracks (WECs) ausfallen, wenn der Wälzkontakt neben der Haupt-Wälzbeanspruchung (p Hz ) noch weitere, sogenannte Zusatzbeanspruchungen ertragen muss. Aktuell wird in der Literatur diskutiert, ob hohe Pressungen zu Beginn des Betriebs alleine ausreichend für eine WEC-Bildung sein können. Zur Klärung dieser Frage werden kürzlich durchgeführte und publizierte WEC-Versuche auf Wälzlagerprüfständen und Tribometern diskutiert, die bei hohen Pressungen und unterschiedlicher überlagerter Wälzkontaktreibung durchgeführt wurden. Bei Versuchen mit relativ niedrigem Reibenergieeintrag mit Axialrillenkugellagern und Zylinderrollenlagern kam es trotz sehr hoher Pressung zu Betriebsbeginn zu keiner WEC-Bildung. An einfachen Tribometern scheint die Kombination von hoher Pressung und Mindestschlupf und somit indirekt ein Mindest-Reibenergieeintrag für eine beschleunigte WEA- Bildung erforderlich zu sein. Schrägkugellager und Pendelrollenlager bildeten bei hoher Pressung und gleichzeitig hohem Reibenergieeintrag frühzeitig WECs aus. Der Ausfallort belegt, dass die Reibenergie und nicht die Hertzsche Pressung hierfür hauptursächlich war. Das unterschiedliche WEC-Ausfallverhalten der Versuche bei hoher Last lässt sich mit einem reibenergetischen WEC-Lebensdauermodell, abgeleitet aus WEC-Versuchen auf den Schaeffler- Prüfständen R4G und FE8, gut nachvollziehen. Schlüsselwörter Wälzlager, Reibung, White Etching Cracks, WEC, White Structure Flaking Rolling bearings can fail in rare cases far ahead the calculated rating life due to White Etching Cracks (WECs). Prerequisites for WEC formation are the so called additional loads (electrical current, high friction, etc.) acting besides the Hertzian rolling contact stresses as the main load. Currently, it is being discussed in the literature whether a high pressure at the beginning of operation can be sufficient to initiate WECs. To address this question, recently performed as well as published WEC tests on bearing test rigs and basic tribometers working with high Hertzian pressures and various levels of internal friction energy are reviewed and discussed. It could be obtained from ball and roller bearing tests that high Hertzian pressures in conjunction with low specific friction energy do not lead to WEC formation. So, evaluating simple tribometer tests, it can be assumed that a combination of high pressure and a minimum proportion of slip (leading also to friction energy ingress) is required to trigger an accelerated WEA formation. In contrast, WECs occur in angular contact ball bearings and spherical roller bearings if they are subjected to high contact pressures and high specific friction energy at the same time. Additionally, the failure location indicates that the friction energy was the main trigger. A frictional WEC lifetime model, which was derived from a large number of bearing tests (FE8, R4G) before, provides an explanation for these different failure behaviors in the WECtests. Keywords Roller bearings, White Etching Crack, WEC, White Etching Area, White Structure Flaking, friction Kurzfassung Abstract * Dr. Jörg Loos Orcid-ID: https: / / orcid.org/ 0000-0001-7596-3205 Dr. Wolfram Kruhöffer Orcid-ID: https: / / orcid.org/ 0000-0002-6848-3040 Jörg Franke Orcid-ID: https: / / orcid.org/ 0000-0002-5227-3620 Schaeffler Technologies AG & Co. KG 91074 Herzogenaurach Daniel Merk Orcid-ID: https: / / orcid.org/ 0000-0002-3845-9146 Toni Blaß Orcid-ID: https: / / orcid.org/ 0000-0002-1286-8144 Schaeffler Technologies AG & Co. KG, 97421 Schweinfurt Phasen (White Etching Areas, WEAs) bestehen größtenteils aus sehr feinkörnigem, kohlenstoffübersättigtem Ferrit. Sie enthalten keine oder sehr kleine Karbide [Hol15]. Drei Erscheinungsformen von WECs können beobachtet werden (siehe Bild 1). Schliffe, in denen WEA Netzwerke dominieren und Risse eher nur flankierend auftreten. Häufig wird auch der umgekehrte Fall beobachtet. Die Risse erscheinen dann stärker ausgeprägt als die WEAs. Des Weiteren findet man in seltenen Fällen auch Strukturen, die an langgestreckte Butterflies erinnern. Der WEC-Entstehungsmechanismus ist bis heute nicht abschließend geklärt. Viele Ermüdungsversuche mit wasserstoffbeladenen Lagern belegen aber, dass bereits ein initial erhöhter Wasserstoffgehalt eine WEC-Bildung initiieren kann [z. B. Veg10, Rue14, Din18]. Außerdem zeigen sie, dass die Schwächung des Gefüges durch Wasserstoff zu Beginn des Einsatzes ausreichend ist. Auf der anderen Seite konnte nachgewiesen werden, dass Wälzlager im Betrieb durch hohe Wälzkontaktreibung oder auch Stromdurchgang atomaren Wasserstoff aufnehmen [Kue15, Ric18, Koh06, Han16, Geg18]. Die Wasserstoffentstehung durch spezielle Korrosionsmechanismen, insbesondere in Anwesenheit von Wasser und einem Gleichstrom, ist ebenfalls gut bekannt [z. B. Ros05]. Der in Bild 2 gezeigte „Wasserstoff-WEC-Pfad“ gibt somit eine mögliche Erklärung, warum WEC-Schäden oft in Anwendungen mit hoher Wälzkontaktreibung, elektrischem Stromdurchgang oder hohem Korrosionsrisiko beobachtet werden. Stark umstritten ist in der Literatur aktuell der genaue metallphysikalische Prozess (spannungsgetriebene Kohlenstoffdiffusion, lokale schwere plastische Verformung, Rekristallisierung, Rissbildung) und die Fragestellung, ob sich die Risse [z. B. Man19] oder die WEAs [z. B. Oez18] zuerst bilden. Nach den Erkenntnissen der allgemeinen Materialforschung an Stählen sind beide Wege möglich [Her12]. Hohe lokale Wasserstoffkonzentrationen fördern die Rissbildung (HEDE, HIC) aber auch lokale plastische Verformungen (HELP) mit anschließender WEA-Bildung. Möglich ist, dass die Betriebsbedingungen (Ver- Wind-Tribology 11 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0024 Bild 1: Erscheinungsformen der WEC-Bildung Bild 2: Erkenntnisstand WEC-Bildung in Wälzlagern (Standard-Wälzlagerstahl) Wind-Tribology 12 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 initiierte WECs bei relativ niedriger spezifischer Reibleistung, z. B. in typischen WEC-Tests auf dem FE8- Prüfgerät, treten ebenfalls stark abhängig von der Schmierstoffformulierung und dem elektrischen Potential auf (siehe Bild 3). Bei sehr hohen spezifischen Reibleistungen ist der Schmierstoffeinfluss wiederum davon abweichend und deutlich geringer. Dies wird deutlich, wenn man z. B. die Ergebnisse gleicher Schmierstoffe vom Micro-Pitting Rig (MPR) mit extrem hohen spezifischen Reibleistungen [Ric19, Gou19, Man19] mit denen vom FE8-Prüfgerät [Ric19] vergleicht. Auch existiert bei der energetischen WEC-Ermüdung nicht mehr zwingend ein Einfluss des elektrischen Potentials (siehe Bild 3). Reibenergetische WEC-Kennzahlen und WEC-Lebensdauermodell Zur Ermittlung einer geeigneten WEC-Beanspruchungskenngröße für die durch Reibung induzierten WEC- Schäden wurden in [Kru16] Wälzlagerversuche mit dem gleichen WEC-kritischen Schmierstoff durchgeführt. Die WEC-Neigung korrelierte hierbei gut mit der kinematischen Reibenergie-Akkumulation e a,kin (siehe Bild 4), bei der die flächenbezogene Reibenergie, mit der ein Oberflächenelement bei Überrollung beaufschlagt wird, ins Verhältnis zur Zeitspanne zwischen zwei Überrollungen gesetzt wird. Des Weiteren hatten bei den Versuchen die relative Schmierfilmdicke ( Λ 0 ) und die Kontaktart (Kugel vs. Rolle, Ring) wesentlichen Einfluss auf die WEC-Neigung. Versuche an Modellprüfständen, verbunden mit Wasserstoffsimulationen zeigten außerdem, dass auch die Größe des Kontaktes eine Rolle spielt [Fra17]. Zur Berücksichtigung dieser Einflüsse wurden der Schmierungskorrekturfaktor φ Λ , der Breitenfaktor k Breite und der Kugelfaktor k Kugel eingeführt [Loo17]. In der Praxis werden viele Anwendungen zwar mit potentiell WEC-kritischen Schmierstoffen betrieben hältnis Wasserstoffkonzentrationen zu Vergleichsspannungen) darüber entscheiden, welcher der beiden Prozesse bestimmend ist, wodurch sich auch die unterschiedlichen WEC-Erscheinungsformen erklären ließen. Einige Forscher arbeiten auch am Nachweis eines alternativen WEC-Entstehungspfads (siehe Bild 2), bei dem elektrische Ströme elektrothermisch [Sce15] oder sehr hohe Spannungen durch Überlast [Sta17], Stoßbelastungen [Bru19] oder Kantenpressungen [Man19] das Gefüge zu Lebensdauerbeginn schwächen und WECs auslösen. WEC Hauptmechanismen WEC-Versuche deuten darauf hin, dass es zwei WEC- Hauptmechanismen gibt. Kleine Gleichströme z. B. infolge elektrostatischer Aufladungen [Loo16] führen zu WEC-Schäden ausschließlich am kathodisch geschalteten Lagerring (siehe Bild 3). Die Lebensdauer ist bei dieser „kathodischen WEC-Ermüdung“ stark abhängig von der verwendeten Schmierstoffformulierung. Bei in der Spitze sehr hohen elektrischen Lagerströmen, wie sie z. B. in umrichtergespeisten E-Maschinen vorkommen können, ist der Einfluss der elektrischen Polung hingegen wesentlich geringer und sogar konträr zur kathodischen WEC-Ermüdung. Hier besitzen die anodisch geschalteten Ringe ein höheres WEC-Risiko [Mik07, Din18]. WEC-Treiber im Fall hoher elektrischer Ströme scheinen die mit hoher Häufigkeit im Schmierspalt auftretenden Blitzentladungen zu sein, bei denen kurzzeitig Oberflächentemperaturen von über 1000 °C auftreten können. Der Einfluss der Schmierstoffformulierung bei der dann auftretenden „energetischen WEC-Ermüdung“ ist außerdem deutlich abweichend von der „kathodischen WEC-Ermüdung“ und wesentlich geringer. Auch bei reibungsinitiierter WEC-Bildung scheint es diese beiden Hauptmechanismen zu geben. Mischreibungs- DOI 10.24053/ TuS-2024-0024 Bild 3: WEC-Hauptmechanismen / Einfluss der elektrischen Polung [Sur14], ohne dass jedoch WEC-Schäden beobachtet wurden. Dies deutet daraufhin, dass WECs nur entstehen, wenn die Reibenergie oder die elektrische Beanspruchung einen Schwellwert überschreiten. Dies bestätigen auch Versuche nach [Loo17], bei denen unterhalb einer kritischen Reibbeanspruchung auch mit einem „WEC-kritischen“ Schmieröl keine WECs mehr auftraten. Die WEC-Dauerfestigkeitsgrenze hing dabei aber auch von der Hertzschen Pressung ab. Um dies bei einer WEC-Risikoabschätzung berücksichtigen zu können, wurde die pressungsgewichtete Reibenergie-Akkumulation e *a,effektiv,pHz-gewichtet als neue Kennzahl für die „Gesamt“-WEC-Neigung - bestehend aus Zusatz- und Hauptbeanspruchung - eingeführt (siehe Bild 4). Nach der vorgestellten wasserstoffbasierten WEC-Schadenshypothese hängt die WEC-Lebensdauer nicht nur von der reibenergetischen Beanspruchung, sondern auch von der WEC-Beanspruchbarkeit des Werkstoffes sowie dem verwendeten Schmierstoff als potentielle Wasserstoffquelle ab. Außerdem muss zwischen energetischer und kathodischer WEC-Ermüdung unterschieden werden, da sich die WEC-Neigung des Werkstoffs, der Lebensdauereinfluss des Schmierungszustandes und der Schmierstoffformulierung dabei stark unterscheiden. Die Ermittlung der in die Lebensdauerformel eingehenden Faktoren für Werkstoff und Schmierstoff erfolgt hierbei experimentell jeweils durch Vergleich mit einem Referenzwerkstoff und Referenzschmierstoff. WEC-Bildung bei hohen Lasten Im FE8-Prüfgerät entstehen WECs in Kugellagern (51212, 7312) nach ähnlich kurzen Laufzeiten wie in dem Axial-Zylinderrollenlager 81212, wenn die Hertzschen Pressungen stark erhöht werden. Beim Axial- Rillenkugellager liegen diese sogar deutlich über 3000 N/ mm 2 (siehe Bild 5), was über die kinematische Reibenergie-Akkumulation gut erklärbar ist, da beim Axial-Rillenkugellager Pressungs- und Gleitgeschwin- Wind-Tribology 13 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0024 Bild 5: Lebensdauern WEC-Versuche abhängig von Lagertyp und Pressung [Loo17] Bild 4: Reibenergetisches WEC-Lebensdauermodell (siehe auch [Kru16]) Wind-Tribology 14 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 se hochbelastet werden und insbesondere, wenn die Lastrichtung nicht passend zur Lagertype gewählt wird. Da bisher nur relativ kleine Wälzlager untersucht wurden, stellt sich auch die Frage, ob sich große Wälzlager abweichend verhalten. Dafür wurden Versuche an für Ermüdungsversuche sehr großen Pendelrollenlagern der Type 22332 (d = 160 mm) durchgeführt. Diese belegen, dass auch große Pendelrollenlager durch WECs infolge sehr hoher Reibung ausfallen können und die Laufzeit zu den von kleinen Lagern abgeleiteten WEC-Lebensdauern passt. Wie Bild 7 verdeutlicht, entstehen die WECs nicht am für die klassische Wälzermüdung kritischeren Innenring, sondern am Außenring. Es ist die digkeitsverläufe gegenläufig sind: Die Pressung (p max ) ist in Laufbahnmitte, die Gleitgeschwindigkeit (v max ) am Laufbahnrand, maximal. Hohe spezifische Reibleistungen („pv“-Werte) ergeben sich bei konstanter Drehzahl deshalb erst bei relativ hohen Pressungen. Der Umstand, dass trotz sehr hoher Pressungen die WEC-Laufzeiten beim Lager 51212 gegenüber den anderen, eher moderat belasteten Lagern, nicht merklich abfallen, spricht gegen die These, dass sehr hohe Pressungen infolge Überlasten oder Kantenpressungen einen zusätzlichen WEC-Trigger darstellen. Versuche von Ruellan [Rue14] an hochbelasteten Schrägkugellagern deuten ebenfalls darauf hin, dass nicht alleine eine hohe Pressung, sondern die beim Schrägkugellager stark lastabhängige Reibung durch Differential- und Bohrschlupf WEC-auslösend ist. Die WECs bildeten sich dort, wo der Reibenergieeintrag am größten war und nicht in den Bereichen sehr hoher Pressungen in Laufbahnmitte (siehe Bild 6). Wurden die Lager mit Wasserstoff aufgeladen, entstanden die WECs hingegen dort, wo die Hertzsche Pressung maximal war, im Bereich der Laufbahnmitte, was die Hypothese stützt, dass der lokal in Bereichen hoher Reibung gebildete Wasserstoff WEC-auslösend ist. Ähnlich wie bei Schrägkugellagern entstehen auch bei Pendelrollenlagern hohe Differentialschlupfe, wenn die- DOI 10.24053/ TuS-2024-0024 Bild 6: Belastungsabhängige Position der WEC- [Rue14] Bild 7: WEC-Versuche an großen Pendelrollenlagern (22332, d m = 250 mm) Wind-Tribology 15 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 Position betroffen, an der die reibenergetische WEC- Beanspruchung (e *a,effektiv,pHz-gewichtet ) am größten ist und nicht in Kontaktmitte, wo die Hertzsche Pressung maximal ist. Es ist aus WEC-Versuchen mit wasserstoffaufgeladenen Wälzlagern auch bekannt, dass der diffusible Wasserstoff bzw. die kritische Zusatzbeanspruchung nur kurze Zeit zu Betriebsbeginn wirken muss. Das liefert eine mögliche Erklärung, warum bei den in [Sta17] durchgeführten Versuchen eine kritische Beanspruchungsphase von ca. 10 Minuten ausreichte, um WECs auszulösen. In dieser kurzen Schädigungsphase sieht hier das Radial- Pendelrollenlager 23024 aufgrund einer reinen Axiallast eine sehr hohe maximale Hertzsche Pressung von 3800 N/ mm 2 , aber auch eine sehr hohe reibenergetische WEC-Beanspruchung (e a,effektiv,pHz-gewichtet ≈ 440 %). Die extrem hohe Axiallast kann zusammen mit der Kombination aus Außenring-Spielpassung und Teilbestückung, die teilweise verwendet wurden, zusätzlich noch zu sehr hohen Zugspannungen im Außenring führen, die möglicherweise auch die WEC-Neigung erhöhen [Lai16]. Zur Klärung, ob wie in [Sta17] vermutet, wirklich nur die Spitzenpressung von 3800 N/ mm 2 zu Betriebsbeginn WEC-auslösend war, wurden Axial-Rillenkugellager im FE8-Prüfgerät bei gleicher maximaler Pressung (p Hz,max = 3800 N/ mm 2 ) und Lastspielzahl (37000), aber sehr geringer reibenergetische WEC-Beanspruchung (e *a,effektiv,pHz-gewichtet ≈ 7 %) zu Laufzeitbeginn geprüft. Obwohl in der anschließenden Ermüdungsphase die Hertzsche Pressung mit p Hz,max = 2350 N/ mm 2 sogar noch etwas größer als beim Pendelrollenlager-Versuch war, fielen die Axial-Rillenkugellager auch nach etwa dreifacher modifizierter Referenzlebensdauer nicht mit WECs aus (siehe Bild 8). Vergleichbare Ergebnisse wurden im FVA-Vorhaben „Wälzlagerlebensdauer-Windgetriebe“ [FVA11] gewonnen. Dort wurden Pressungen von 3000 N/ mm 2 zu Beginn des Versuches als „Überlast“ aufgebracht - was ebenso nicht ausreichte, um Frühausfälle zu generieren. Wie die Axial-Rillenkugellager-Versuche im FE8- Prüfgerät wurden die Versuche im FVA-Vorhaben mit Zylinderrollenlagern bei geringer reibenergetische WEC-Beanspruchung (e *a,effektiv,pHz-gewichtet < 10 %) durchgeführt. Auch Manieri [Man19] stellt die Hypothese auf, dass sehr hohe Pressungen alleinig WECs auslösen können. In seinen Versuchen traten WECs nur auf, wenn nicht profilierte Prüflinge mit hohen Kantenpressungen (p Hz,max,Kante > 4000 N/ mm 2 ) eingesetzt wurden. Die Prüflinge ohne Kantenpressungen mit dadurch deutlich geringeren maximalen Pressungen fielen zwar relativ früh, aber ohne WECs aus. Es ist zu beobachten, dass alle Versuche jedoch mit mindestens 5 % Schlupf durchgeführt wurden. Dies, in Kombination mit einer hohen Kantenpressung, mit für Ermüdungsversuche typischen hohen Überrollfrequenzen und Mischreibung, führt zu extrem hohen reibenergetischen WEC-Beanspruchungen (e *a,effektiv,pHz-gewichtet > 1000 %), welche immer deutlich größer waren als in allen gängigen WEC-Wälzlagertests (z. B. im R4G-WEC-Test [Loo16b] oder FE8- WEC-Tests [Loo17]). Gemäß [Bru19] ergibt sich eine beschleunigte WEA-Bildung in 2-Scheibentests auch nur bei gleichzeitigem Auftreten einer hohen Pressung und eines hohen Schlupfes (p Hz > 2,4 GPa und Schlupf > 5 %). Bisher ist somit nicht zweifelsfrei geklärt, ob sehr hohe Pressungen auch ohne hohe Reibung WEC-auslösend sein können. Die vorgestellten Ausfälle sind alle über den „Wasserstoff-WEC-Pfad“ (siehe Bild 2) erklärbar. Dies wird insbesondere bestätigt, wenn man die Versuchspunkte (WEC-Beanspruchung, Laufzeiten) in die Wöhlerlinie für energetische WEC-Ermüdung einträgt (siehe Bild 9). Auch wenn vereinzelt Abweichungen zwischen WEC-Laufzeitprognose (gestrichelte Linie) DOI 10.24053/ TuS-2024-0024 Bild 8: Vergleich „Hochlast“-WEC-Versuche mit 23024 und FE8-Versuche mit 51212 Wind-Tribology 16 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 modifizierte Referenzlebensdauer ohne Schaden. Dies passt auch zu Versuchen aus dem FVA-Vorhaben 541, in dem Pressungen von 3000 N/ mm 2 , zu Beginn als „Überlast“ aufgebracht, nicht ausreichten, um an Zylinderrollenlagern mit geringer reibenergetischer Beanspruchung Frühausfälle zu erzeugen. Untersuchungen zur WEA-/ WEC-Bildung werden teilweise auch an einfachen Tribometern (2-Scheiben, MPR) durchgeführt [Man19, Bru19]. Die maximalen Hertzschen Pressungen und Gleitgeschwindigkeiten sind bei diesen Versuchen immer sehr hoch. Dort kann deshalb nicht sauber zwischen hoher Pressung und hoher Reibenergie als potentiellen WEC-Trigger getrennt werden. Aus den Tribometerversuchen von Bruce [Bru19] ergibt sich eine beschleunigte WEA-Bildung jedoch nur bei gleichzeitigem Auftreten einer hohen Pressung und eines hohen Schlupfes (p Hz > 2,4 GPa und Schlupf > 5 %). Dies und die Tatsache, dass sich die Laufzeitunterschiede der vorgestellten WEC-Versuche mit Hilfe des WEC-Lebensdaueransatzes für energetische WEC- Ermüdung sehr gut erklären lassen, bekräftigten die Hypothese, dass hohe Pressungen nur dann zu WECs führen, wenn als Zusatzbeanspruchung zur Wasserstofffreisetzung ein WEC-Trigger wirksam ist: Stromdurchgang, Korrosion und/ oder ein hoher spezifischer Reibenergieeintrag. Bezeichnungen B h Versuchslaufzeit in Stunden e a, effektiv effektive Reibenergie-Akkumulation (e a,kin , Λ 0 , Kontaktbreite) e a,effektiv,pHz,gewichtet pressungsgewichtete Reibenergie-Akkumulation (e a,effektiv , p Hz ) e a,kin kinematische Reibenergie-Akkumulation, („pv“-Wert, Regenerationszeit) und Versuchslaufzeit erkennbar sind, so lassen sich die Laufzeitunterschiede zwischen den Versuchen gut mit Hilfe des reibenergetischen WEC-Lebensdauermodells begründen. Zusammenfassung Sowohl bei niedrigen als auch bei hohen Hertzschen Pressungen kann es in Wälzlagern zur WEC-Bildung abhängig von der Höhe der Zusatzbeanspruchung (Reibung, Stromdurchgang, …), des verwendeten Schmierstoffs (Additivierung, Wassergehalt, …) und Werkstoffs kommen. Bei hohen Pressungen kann die Gleitreibung z. B. infolge Differential- oder Bohrschlupf einen kritischen WEC-Trigger darstellen. Da dieser Reibanteil stark lagertypabhängig ist, hängen im FE8-Prüfgerät die WEC-Lebensdauern mehr vom Lagertyp als von der maximalen Hertzschen Pressung ab [Loo17]. Die WECs bilden sich an hochbelasteten Schrägkugellagern demzufolge auch dort, wo der Reibenergieeintrag durch den Schlupf maximal ist und nicht in den Bereichen maximaler Pressung in Laufbahnmitte [Rue14]. Gleiches Verhalten zeigten auch große Pendelrollenlager unter hoher Belastung und kinematisch ungünstiger Belastungsrichtung. WECs entstanden am Außenring mit niedrigerer maximaler Hertzschen Pressung, aber deutlich höherer reibenergetischer WEC-Beanspruchung als am Innenring. Dass eine Zusatzbeanspruchung z. B. in Form eines sehr hohen Reibenergieeintrags erforderlich und eine hohe Pressung zu Betriebsbeginn alleine nicht ausreichend ist, ließ sich an Axial-Rillenkugellagern zeigen. Obwohl die Lager zu Versuchsbeginn mit einer sehr hohen Pressung von 3800 N/ mm 2 beaufschlagt wurden (bei bewusst sehr niedriger Reibbeanspruchung), kam es zu keiner WEC-Bildung. Die Lager erreichten die dreifache DOI 10.24053/ TuS-2024-0024 Bild 9: Vergleich WEC-Lebensdauerprognose mit Versuchslaufzeiten f Ü Überrollfrequenz k Breite Korrekturfaktor WEC-Lebensdauereinfluss Kontaktbreite k Kugel Korrekturfaktor WEC-Lebensdauereinfluss Kugel L h, WEC WEC-Lebensdauer in Stunden N Lastspielzahl N Ref Lastspielzahl des Referenzsystems bei Referenzbedingungen N WEC Median der Lastspielzahl bis zum WEC-Ausfall p Hauptbeanspruchung WEC-Lebensdauerexponent für Hertzsche Pressung p Hz,max maximale Hertzsche Pressung p Zusatzbeanspruchung WEC-Lebensdauerexponent für Zusatzbeanspruchung t Reg Zeit zwischen 2 Überrollungen (Regenerationszeit) v Gleitgeschwindigkeit Λ 0 Relative Schmierfilmdicke (zentrale Filmdicke / Summenrauheit) φ Schmierstoff Korrekturfaktor WEC-Lebensdauereinfluss Schmierstoff φ Werkstoff Korrekturfaktor WEC-Lebensdauereinfluss Werkstoff φ Λ 0 Korrekturfaktor WEC-Lebensdauereinfluss rel. Schmierfilmdicke Λ 0 κ Viskositätsverhältnis Literatur [Bla17] Blass, T. et al.: Influence of Material and Heat Treatment on the Formation of WECs on Test Rig FE8, Advances in Steel Technologies for Rolling Bearings STP 1580, ASTM 2017 [Bru19] Bruce, T. et al.: Threshold Maps for Inclusion-Initiated Micro-Cracks and White Etching Areas in Bearing Steel: The Role of Impact Loading and Surface Sliding, Tribology Letters (2018) 66: 111 [Din18] Dinter, R.et al.: Formation and Detection of Pre- Stages of White Etching Cracks (WEC), 3. VDI- Fachkonferenz Schadensmechanismen an Lagern, Aachen 2018 [Eva16] Evans, M.-H.: An updated review: white etching cracks (WECs) and axial cracks in wind turbine gearbox bearings, Material Science and Technology 2016 [Fra17] Franke, J. et al.: Untersuchung der Übertragbarkeit von WEC-Wälzlagerversuchen auf Modellprüfstände, Antriebstechnisches Kolloquium, ATK, Aachen 2017 [FVA11] N. N.: Wälzlagerlebensdauer-Windgetriebe, Berücksichtigung von Betriebszuständen, Sonderereignissen und Überlasten bei der Berechnung der Wälzlager-Lebensdauer in Windenergieanlagen und Großgetrieben, FVA 541 I, FVA-Heft Nr. 967, 2011 [Geg18] Gegner, J et al.: Wälzlagerschäden mit weiß anätzenden Rissen abseits der Windenergie, 3. VDI-Fachkonferenz Schadensmechanismen an Lagern, Aachen Juli 2018 [Gou19] Gould, B. et al.: The Effect of Lubricant Composition on White Etching Crack Failures, Tribology Letters (2019) 67: 7 [Han16] Han, B. et al.: In Situ Detection of Hydrogen Uptake from Lubricated Rubbing Contacts, Tribology Online 11 450-454, 2016 [Her12] Hertzberg, R. W.: Deformation and fracture mechanics of engineering materials. Wiley, 2012. [Hol15] Holweger, W. et al.: White Etching Crack Root Cause Investigations,Tribology Transactions 2015, 58(1), pp 59-69 [Mik07] Mikami, H. et. al.: Influence of electrical current on bearing flaking life, SAE Technical Paper 2007-01- 0113, SAE International, Warrendale, USA 2007 [Koh06] Kohara, M. et al.: Study on Mechanism of Hydrogen Generation from Lubricants NTN Corporation, Tribology Transactions, 49: 53-60, 2006, NTN [Kru16] Kruhöffer, W. Loos, J.: WEC Formation in Rolling Bearings under Mixed Friction: Influences and „Friction Energy Accumulation“ as Indicator, Tribology Transaction 2016 [Kue15] Kürten, D. R.: Einfluss der tribochemischen Schmierstoffoxidation auf die wasserstoffinduzierte Wälzkontaktermüdung, Fraunhofer IWM Forschungsberichte Band 7, Dissertation 2015 [Lai16] Lai, J., Stadler, K.: Investigation on the mechanisms of white etching crack (WEC) formation in rolling contact fatigue and identification of a root cause for bearing premature failure, Wear 364-365 (2016), p. 244-256 [Loo16a]Loos J., et al.: Influence of Currents from Electrostatic Charges on WEC Formation in Rolling Bearings, Tribology Transactions 2016, Volume 59, 865-875 [Loo16b]Loos, J. et al.: Factors increasing the risk of WEC - formation in large-size bearings with full fluid film lubrication, Bearing World conference Hannover 2016 [Loo17] Loos, J. et al.: Berechnungsansätze für die WEC- Neigung unter reibenergetischer Wälzbeanspruchung, 12. VDI-Tagung Gleit- und Wälzlagerungen 2017, VDI-Berichte 2308, S. 231 [Man19] Manieri, F. et al.: The origins of white etching cracks and their significance to rolling bearing failures, International Journal of Fatigue, Volume 120, March 2019, Pages 107-133 [Oez18] Oezel, M. et al.: Formation of white etching areas in SAE 52100 bearing steel under rolling contact fatigue - Influence of diffusible hydrogen, wear 414- 415 (2018) 352-365 [Ric18] Richardson, A. D. et al.: Thermal Desorption Analysis of Hydrogen in Non hydrogen Charged Rolling Contact Fatigue Tested 100Cr6 Steel, Tribology Letters 66: 4, 2018 Wind-Tribology 17 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0024 Wind-Tribology 18 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 [Sur14] Surborg, H.: Einfluss von Grundölen und Additiven auf die Bildung von WEC in Wälzlagern, Dissertation Universität Magdeburg, Aachen, 2014 [Sta17] Stadler, K. et al.: Short term heavy loads - one cause for premature bearing failures and white etching cracks, GfT Tagung Göttingen 2017 [Veg10] Vegter, R. H. Slycke, J. T.: The Role of Hydrogen on Rolling Contact Fatigue Response of Rolling Element Bearings, Journal of ASTM international, Vol. 7, No. 2 Initial publication: 67/ 1 [Ric19] Richardson, A. D. et al.: The effect of over-based calcium sulfonate detergent additives on white etching crack (WEC) formation in rolling contact fatigue tested 100Cr6 steel, Tribology International 2019 [Ros05] Rossmeisl, J. et al.: Electrolysis of water on (oxidized metal surfaces, Chemical Physics 319, p178-184, 3005 [Rue14] Ruellan, A.: Tribological analysis of White Etching Crack (WEC) failures in Rolling Element Bearings, L’Institut National des Sciences Appliquées de Lyon, Thesis 2014 [Sce15] Scepanskis, M. et al.: The Numerical Model of Electrothermal Deformations of Carbides in Bearing Steel as Possible Cause of White Etching Cracks Initiation, Tribology Letters, vol. 59, 2015 DOI 10.24053/ TuS-2024-0024 Introduction With its higher efficiency than solar panels, wind power is a fast-growing worldwide industry, with over 650GW of capacity installed all over the world now. 1.9 GW of freshly installed offshore wind capacity brought the total German fleet to 56.1 GW in 2021. This makes up over 28.000 turbines for now with number rising. Wind-Tribology 19 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0025 Simplified tribological approach for predesign of wind turbine bearing cases, combined with model test investigation Ivan Grozev, Sagar Dalal, Nazlim Bagcivan, Serhan Bastuerk, Christian Lueffe, Thomas Stahl* Dieser Beitrag wurde im Rahmen der 63. Tribologie-Fachtagung 2022 der Gesellschaft für Tribologie (GfT) eingereicht. Mit dem Ziel, unsere Vorgehensweise des Produktentwurfs zu optimieren und durch den Übergang von verschwenderischen Feldtests zu schlanken tribologischen Modelversuchen die Nachhaltigkeit des Endprodukts zu erhöhen sowie ein wirkungsvolles und umsetzbares Systemranking für Lagergehäuse von Windkraftanlagen zu erreichen, befassten wir uns näher mit den folgenden gängigen Berechnungsansätzen: • Plint’s and Alliston-Greiner’s Energy Pulse (EP) • Matveesky’s friction power intensity (FPI) • Transmitted Energy (TE) Mit dem Konzept der übertragenen Reibenergie (Ef) erreichten wir zufriedenstellende Ergebnisse. Des Weiteren stellten wir fest, dass Zwei-Scheiben- Tribometer verwendet werden können, um einen Gleitvorgang von 100 % zu simulieren. Dies erleichtert den qualitativen und quantitativen tribologischen Vergleich von Systemen mit unterschiedlichen Schlupfraten durch die Verwendung von nur einer Maschinentype, was die Messsicherheit erhöht. Schlüsselwörter Produktentwurf, Nachhaltigkeit, Reibenergie, Reibintensität, Energiepuls, Zwei-Scheiben Tribometer, Tribometer, Tribometrie, Tribolabor, Tribologie With the purpose of improving our predesign approach and increasing the end-product sustainability by going from resource wasting field testing to slender tribological model testing, we looked closer to commonly relevant calculation approaches as • Plint’s and Alliston-Greiner’s Energy Pulse (EP) • Matveesky’s friction power intensity (FPI) • Transmitted Energy (TE) to achieve fast and feasible tribo-system ranking for wind turbine bearing cases. With the concept of transmitted friction energy (Ef), satisfactory results were reached. We learned that 2-Disc tribometer may be used to simulate 100 % sliding, making comparison of different slippage rate easier by only using one tribometer, thus reducing measurement inaccuracies caused by different type of machines (e.g., pin on disc) or operator’s influence. Keywords Product predesign, sustainability, transmitted friction energy, friction power intensity, energy pulse, twodisc machine, tribometer, tribometry Kurzfassung Abstract * Ivan Grozev 1 (federführender Autor) Sagar Dalal 2 Dr. Serhan Bastuerk 1 Dr. Nazlim Bagcivan 1 Christian Lueffe 2 Dr. Thomas Stahl 2 1 Schaeffler Technologies & Co.KG, 91074 Herzogenaurach 2 Schaeffler Technologies & Co.KG, 97421 Schweinfurt Wind-Tribology 20 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 • Pin type cages that consist of two side washers and a bolt • Stees Segment Cages (SSCs) optimized for clean and easy production. Within our investigation, we looked closer at the Archard’s wear law, the Matveesky’s friction power intensity, Plint’s and Alliston-Greiner’s energy pulse approach and considered that the concept of transmitted friction energy (Ef) appears to rank the investigated cage systems in a clear and relatively precise way. Compared to a system description, mainly based on the product of load and relative velocity (p*v), the transmitted energy approach is much more accurate, quickly pinpointing possible system performance challenges and thus directing a pre-design process more precise. Additionally, we learned that the two-disc tribometer may be used to technically simulate complete sliding, making comparison of different slippage rate easier by only using one tribometer, thus reducing measurement inaccuracies caused by different type of machines (e.g., pin on disc) or operator’s influence. Simplified tribological approach for predesign of wind turbine bearing cases, combined with model test investigation At the beginning of the 20 th century, the American essayist and naturalist John Burroughs said: “The fuel in Earth will be exhausted a thousand or more years, and its mineral wealth, but man will find substitutes for these in the winds, the waves, the sun’s heat, and so forth.” DOI 10.24053/ TuS-2024-0025 With wind energy pros as a clean and renewable source of energy, relatively low operating costs, effectiveness and decreasing prices due to technological advancements and increased demand, there is little surprise that a global player as the Schaeffler Group concentrates in various solutions specially designed for wind energy turbines. For the blade adjustments, the Schaeffler Group offers plain bushes, spherical plain bearings, deep grove ball bearings, planet gears and output shafts. For the rotor shaft - the single bearing concept and sensor monitoring. In the gearbox there is the planet carrier, the planet gear, the hollow shaft, the intermediate output shaft. There are solutions for the generator and wind tracking. With such a product variety, there appears naturally the need of clear design vision and the possibility of prompt and effective model testing. In this work we try to answer the following questions: • Is it possible to predict, due to design qualities which bearing cage is more vulnerable to wear, this reducing the tendency of overengineering, and on the other hand, pile up on robustness for bearings that tend to need it more? • Can a two-disc model test deliver plausible answers for pure sliding contacts with the purpose of substituting materials in order to optimize and concurrently reduce costs of the production process? We concentrated on four different types of bearing cages for the latter investigation: • Inner ring guided FBP thrust ball bearing cages • JP3 type roller-guided sheet metal cages Figure 1: Overview of Schaeffler Group product applications for wind turbines Wind-Tribology 21 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 This statement is as valid today as more than a hundred years ago. 1.9GW of freshly installed offshore wind capacity brought the total German fleet to 56.1GW in 2021 [1]. This makes up over 28.000 turbines for now with numbers raising. The latter message is taken from the study commissioned by the country’s wind association BWE. With wind energy pros as a clean and renewable source of power, low operating cost (1-2c per KWh), effectiveness and decreasing prices due to technological advancements and increased demand, there is little surprise that a global technology player as the Schaeffler Group concentrates on various solutions specially designed for wind energy turbines (see Figure 1 below): For the blade adjustments we offer plain bushes, spherical plain bearings, deep groove ball bearings, planet gears and output shafts. For the rotor shaft, we have the single bearing concept and sensor monitoring. In the gearbox there is the planet carrier, the planet gear, the hollow shaft, the intermediate output shaft. There are solutions for the generator and wind tracking. Additionally, the Schaeffler Group offers condition monitoring, simulation, calculation and testing. With such a product variety, there appears naturally the need of clear design vision and the possibility of prompt and effective model testing. We tried to answer, via tribological means the following questions: • Is it possible to predict, due to design qualities which bearing cage is more vulnerable to wear, thus reducing the tendency of overengineering, and on the other hand, pile up on robustness for bearings that tend to need it more? • Can a two-disc model test deliver plausible answers for pure sliding contacts that correlate with our field experience with the purpose of substituting materials in order to optimize and concurrently reduce costs of our production process? We concentrated on four different types of bearing cages for our investigation: • Inner ring guided FPB thrust ball bearing cages, which are used successfully for many years. • JP3 cages, which are roller-guided sheet metal cages for large size tapered roller bearings. Generally, the validation of the JP3-design logic is based on track record with over 150 different JP type cages and over 6.000 cages in operation. • Pin type cages that consist of two side washers and a bolt. To achieve a higher wear resistance the bolt is heat treated and additionally coated. • Steel Segment Cages, or SSCs, that are optimized for clean and easy production. The design allows a flexible scaling from midrange wind bearings up to the largest of applications. The SSC cage design is currently available for double row tapered roller bearings (TRBs). During the first part of our tribological study, we were interested if a simplified calculation, based on compressed theory may correlate our field experience. Our purpose was to give more decisiveness and clearness of a pre-design process, based on facts and numbers, simultaneously saving the time and resources needed for simulation. We looked closely at the work done, due to friction as a measure of wear, based on four different approaches: • Archard’s wear law [2] - there are some limitations, including the fact that the law takes into consideration only the softer of materials and doesn’t contain any properties characterizing adhesion: • Matveesky’s friction power intensity (FPI) [3] - here we had the problem that it defines only the rate of energy generation and doesn’t take contact time into consideration. The normal load and the apparent area of contact in the FPI-equation can be alternatively substituted with the surface pressure p[N/ mm 2 ], resulting in: Within the Schaeffler Group we have gathered empirical knowledge for functional area FPI-values of certain coating types, and we use the latter equation for model test designing purposes. The Q f value may be considered a cross-contact energy and derived from our experience correlates well with the Archard’s wear coeff. k [mm 3 / N] at the end of the model tests. • Plint’s and Alliston-Greiner’s energy pulse approach (EP) [4] - this is a revised version of the FPI that can be “regarded as an incremental contribution to wear or surface damage in contact. Sum of Eps can be used as a measure of total wear” [5]. The energy pulse is only applicable to the type of contact in which the contact point moves relative to the two surfaces, and has obtained encouraging values, which can be used to describe the teeth gear [x]. V = Q f = . = μ [W/ mm²] Q f = = μ [ / ² ] = [mm³] . DOI 10.24053/ TuS-2024-0025 Wind-Tribology 22 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 well. Taking the compressed energy that can be converted to temperature into consideration, we have set a critical border prospectively at 1000J per minute. Still this assumption needs to be further validated. Based on discussions with the application engineering and after looking at different contributing factors, the 100 % ranking of various cage designs is relatively difficult. Still the above results and the approximate classification are plausible, and this simple ranking shows good potential in pre-design approach, pinpointing optimization possibilities and where to look for them. The next important question of our study was if a twodisc model test can deliver plausible answers for pure sliding contacts that correspond to our field experience. The reason for not choosing a pin on disc test was that Schaeffler Group has gathered a considerable experience with two-disc tribometer in the past, even accomplishing to reproduce a very similar wear mechanism close to what we are used to see in field conditions. Additionally, we had a better representation of the contact area geometry, due to the flexibility a two-disc system can offer. On top, a rolling / sliding wear tester is one of the most popular machines for investigating wear and friction of systems under rolling, sliding or a combination of both conditions. We chose a two-disc tribometer by Optimol Instruments Muenchen. The machine operates tow discs, fixed to two parallel shafts, pressed against each other. Driven by a motor through a train of gear, the specimens are rotating along with the shafts. The rotating speed can be controlled, so that when the linear speeds of the two wheels are equal at the contact point (V 1 = V 2 ), a pure DOI 10.24053/ TuS-2024-0025 The transit times for the contact areas are: What we understood from the above approaches is that roughly kinematics, surface pressure, contact geometry and thereof derived contact ratios, velocity, temperature inand out-flow, lubrication and its regime, running time, material and surface properties play a decisive role in the tribological performance. We assumed, that a friction coefficient is closely related to the contact energy, which is a product of the above listed initial tribological system properties and can be regarded as a resulting resistance value needed to propagate a kinematic motion. Based on the latter recognition, we decided to use, we called it, the transmitted friction energy (E f ) as a comparison guidance in order to evaluate the studied cage types (see Table 1 below) The transmitted energy approach appeared to describe and thus rank the cage-systems more precisely, based on and correlating with our field experience, and give a clearer idea about detecting possible difficulties in a predesign stage than the generally assumed and largely used (p*v) [kg/ s 3 ]. Additionally, the E f - values give a hint if there is a trigger possibility for surface tribochemical processes as Ep = . = μ [J/ mm²] = 2 and = 2 . . ( ) E f = [ ] = μ Cage type FBP JP3 Pin type SSC Load (sta c force) in N 2466.66 384.36 2991.40 1.09 1.5 Contact pressure (static load) in N/ mm² 2.9 132.7 129.3 0.1 12.4 Rela ve velocity in m/ s at n=min -1 0.05 0.046 0.017 0.05 0.05 Cage assembly Rollers inserted directly Rollers inserted via hea$ng procedure Pins inserted through rollers, welded, and threaded at both ends Combined segments Circumferen al usage in % 85 92 95 92 Size of cage in mm 500 - 2250 1000 - 3500 1500 - 3500 1000 - 2500 μ 0.4 0.2 0.3 0.2 E f in J at n=min -1 2960 212 915 0,65 0,90 Table 1: Cage system ranking based on E f and rolling contact is achieved. When V 1 and V 2 are different and both wheels are rotating, a combined rolling-sliding takes place [6]. We chose discs with elliptical contact areas as follows R 1X = 40 mm, R 2X = 40 mm; R 1Y = 320 mm, R 2Y = ∞, or simply disc with crowned radius versa a cylindrical disc (see the contact simulation below). Wind-Tribology 23 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0025 Table 2: Average weight loss of diff. material systems after 2-disc model test Figure 2: Hertzian pressure contact simulation We decided to apply minimum of 2 N on the machine, resulting in 150.8 N/ mm 2 surface pressure, which fits quite well with contact pressures in large size real applications [7]. The precise testing conditions and the specific materials cannot be named for apparent reasons, but the result of the study is to be seen in the graph below. Wind-Tribology 24 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 with different slippage easier, by using only one type of machine and not a combination of e.g., pin on disc and two-disc tribometers. References [1] German Wind Energy Association Report 2021 [2] Archard J.F. “Contact and Rubbing of Flat Surface”. Journal of Applied Physics 1953 24 (8): 981-988. doi: 10.1063/ 1.1721448 [2] Archard J.F., Hirst W. “The Wear of Metals under Unlubricated Conditions” 1956-08-02 Proceedings of the Royal Society. A-236 (1206): 297-410. doi: 10.1098/ rspa.1956.0144 [3] Matveesky R.M. “The critical temperature of oil with point and line contact machines”. Trans. ASME. 1965; 87: 754. doi: 10.1115/ 1.3650672 [4] Plint M.A., Alliston-Greiner A.F. “The energy pulse: A new wear criterion and its relevance to wear in gear teeth and automotive engine valve trains”. 1996 Lubrication Science / Volume 8, Issue 3: 223-251. doi: 10.1002/ ls.3010080303 [5] Plint G. Cambridge University Tribology Course 2015, Slides [6] Dalal S. “Wear Investigation of Large Bearing Cages for Wind Turbines”. 2021 M.Sc. Thesis TU Bergakademie Freiberg / Schaeffler Group [7] Kock S., Jacobs G., Bosse D. “Determination of Wind Turbine Main Bearing Load Distribution”. Journal of Physics: Conference Series - 1222 (2019) 012030: 2-6. doi: 10.1088/ 1742-6596/ 1222/ 1/ 012030 Initial publication: 70/ 2 DOI 10.24053/ TuS-2024-0025 The latter represents a number of tested systems with measured weight loss after the test. At least four test pro System were performed. The bar shows the average value of the single measurements. The orange bar represents always one and the same counterpart material. The results correlate to a larger extent with our field experience and observations thus proving that via careful test design and basic system understanding, a complete sliding can be plausibly simulated via a two-disc tribometer. This is true for a roller vs. cage contact but must be still validated for a shaft vs. bearing bushing combination. This realization opens new possibilities in tribological model testing methodology, as a complete sliding is usually simulated on a pin on disc machine. For systems where a simulated slippage ratio varies, performing model testing on only one machine may prove beneficial. The rough but essential lessons learned from this study is that system tribological properties can be promptly and efficiently described via a transmitted friction energy, taking into consideration mainly the friction coefficient, the static load and the relative velocity. Compared to a system, description, mainly based on the product of load and relative velocity (p*v), the transmitted energy approach (E f ) is much more accurate, quickly pinpointing possible system performance challenges and thus directing the pre-design process more precise. Additionally, we learned that two-disc tribometer may be used to technically simulate complete sliding, making tribological model test approach for comparison of systems Wind-Tribology 25 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 Further information and registration at www.tae.de/ go/ tribologie Attend our seminars, courses and conferences. Friction, wear and lubrication Lubricants and operating fluids Lubrication technology Lubricated machine elements A large part of our seminars is supported by the Ministry of Economic Affairs, Labour, and Housing of Baden-Württemberg with funds from the European Social Fund. Benefit from the ESF course funding and secure up to a 70 % subsidy on your participation fee. All information on eligibility for funding can be found at www.tae.de/ foerdermoeglichkeiten Tribology, friction, wear and lubrication Up to 70 % subsidy possible Wind-Tribology 26 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0026 Feasibility study on preloaded flexible conical plain bearings as wind turbine main bearings Jan Euler, Georg Jacobs, Timm Jakobs, Julian Röder* submitted: 16.10.2023 accepted: 17.02.2024 (peer-review) Presented at the GfT Conference 2023 * Jan Euler, M. Sc. Prof. Dr.-Ing. Georg Jacobs Timm Jakobs, M. Sc. Julian Röder, M. Sc. CWD RWTH Aachen 52074 Aachen Windenergieanlagen (WEA) sind eine Schlüsseltechnologie auf dem Weg zu einer kohlenstoffneutralen Energieerzeugung. Derzeit entfallen etwa 30 % der Windenergie-Stromgestehungskosten auf Wartung und Instandhaltung. Eine kritische WEA-Komponente ist das Hauptlager, das innerhalb der üblichen Anlagenlebensdauer von 20 Jahren eine Ausfallwahrscheinlichkeit von 15 bis 30 % aufweist. Der Austausch der Hauptlager ist eine teure Wartungsmaßnahme, da ein externer Kran benötigt wird (~ 250.000 $ pro Tag bei Offshore-WEA). Eine in der Windindustrie diskutierte Maßnahme zur Reduktion der Wartungskosten ist die Verwendung von segmentierten Gleitlagern. Diese Art von Lagern ermöglicht den Austausch von beschädigten Segmenten, ohne, dass der gesamte Antriebsstrang demontiert werden muss. Deshalb hat der Chair for Wind Power Drives (CWD) ein segmentiertes konisches Gleitlagerkonzept für den Einsatz als WEA- Hauptlager entwickelt, getestet und validiert. Zur Ausbildung eines tragfähigen Schmierfilms, benötigen herkömmliche Gleitlager einen konvergenten Kurzfassung Schmierspalt und daher ein Mindestspiel. Dieses Mindestspiel wirkt sich negativ auf die Führungsgenauigkeit aus, da sich die Welle innerhalb der Grenzen des Spiels frei bewegen kann. Mit der zunehmenden Integration von WEA-Antriebskonzepten müssen WEA- Hauptlager höhere Anforderungen an die Führungsgenauigkeit der Hauptwelle erfüllen. Daraus ergibt sich für Gleitlager als Rotorhauptlager eine grundlegende konstruktive Herausforderung. Ein Ansatz zur Erhöhung der Führungsgenauigkeit bei Wälzlagern ist die Verwendung vorgespannter Kegelrollenlager. Bei herkömmlichen Gleitlagern ist eine Vorspannung nicht möglich. In dieser Arbeit wird ein vorgespanntes flexibles konisches Gleitlager als WEA-Hauptlager vorgestellt und bewertet werden. Schlüsselwörter Windenergie, Gleitlager, Hauptlager, Führungsgenauigkeit, Vorspannung Wind turbines (WT) are a key technology towards a carbon neutral energy production worldwide. Currently about 30 % of the Levelized Cost of Electricity (LCoE) consist of service and maintenance. Critical components are the main roller bearings which have a failure probability between 15 and 30 % within 20 years. Main bearing replacements are expensive maintenance procedures because an external crane is needed which costs about $250.000 per day for offshore WTs. Hence, a discussed countermeasure within the wind industry is to use segmented plain bearings in future WTs. These Abstract types of bearings allow an up-tower sub-component wise replacement of faulty parts without the need to dismantle the whole drivetrain. Therefore, the Chair for Wind Power Drives (CWD) developed, tested and validated a segmented conical plain bearing concept for the use as a main bearing for WTs. To function properly common plain bearings need a minimum clearance to allow the formation of a convergent lubrication gab. This initial clearance negatively influences run-out, as the shaft can move freely within 1 Introduction The European Union’s goal is to produce 40 % of its energy through renewables by 2030 [1]. In order to spur on the set up of more wind turbines (WT) the costs need to fall further. The main bearing of WTs is a crucial component regarding maintenance cost. Main bearings suffer from comparatively high failure probabilities of up to 30 % [2]. Currently mostly rolling bearings are commercially available as WT main bearings. In order to exchange a failed rolling main bearing the drivetrain of the WT needs to be disassembled. The exchange of these bearings results in high costs (about $250.000 per day for offshore WTs), as expensive equipment such as cranes or special maintenance vessels are required. Segmented plain bearings as main bearings for WTs are one possible solution to replace the error prone roller bearings. Segmented plain bearings can potentially be repaired up-tower as individual segments can be exchanged by hand or with the use of on-board cranes. One such segmented plain bearing concept was developed and validated at the CWD in the course of the WEA-GLiTS research project [3, 4]. The so-called FlexPad bearing was designed and validated as a main bearing for the Vestas V52 WT (750 kW). The flexible design allows the segments to follow the movement of the shaft, while maintaining parallel surfaces between the sliding segments and the shaft. This behaviour allows large areas for pressure build-up and prevents edge wear [3, 4]. The fundamental design of the FlexPad is shown in Figure 1. The design is characterized by the following key parameters: The cone shape for both bearing halves is determined by their angle of inclination α and the respective inner D i and outer diameter D o . The shape of the sliding segments (pads) and their support structure (arms) is mostly determined by their respective thickness (s pad and s arm ). Moreover, the flexibility of the concept is also characterized by the position x Groove and depth of the groove t groove in the arm. When transferring the concept towards a market relevant scale, changing requirements to the bearing’s performance need to be considered. One such requirement for main bearings is to limit the shafts movement to a specified maximum run-out. For increasing rated power there is a trend towards more integrated drivetrain concepts with high torque density. Typically, there is a high mechanical integration of main bearing, gearbox and generator [6-9]. Due to the integration of main bearing and gearbox the run-out of the main shaft is directly influencing the first gear stage. If the run-out is to large, this would negatively influence the gearbox performance. Common plain bearings are designed with nominal clearances of 0.3 ‰ to 3.5 ‰ of their nominal diameter [10-15]. Within the clearance the shaft is free in its radial movement. The requirement of integrated drivetrains for low run-out results in a need for high stiffness of the main bearing. Especially for the FlexPad concept this constitutes a challenge, as the flexibility of the support structure inherently results in an increased possible shaft run-out. In this work a preloaded FlexPad bearing is investigated and its performance is discussed for different clearance values under production load conditions. Wind-Tribology 27 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0026 the limits of the clearance. As WT drivetrain concepts become more integrated, main bearings for WTs need to fulfill higher requirements regarding the allowable run-out of the main shaft. Therefore, a fundamental design challenge arises for plain bearings as rotor main bearings. One approach to reduce run-out for roller main bearings is to use preloaded tapered roller bearings. For common plain bearings however preloading is not possible. However, the concept of preloading was successfully transferred to the flexible conical plain bearing concept developed at the CWD and the main shaft run-out severely reduced. In this work the feasibility of a preloaded flexible conical plain bearing as a WT rotor main bearing is evaluated and the advantages and disadvantages contrasted. Keywords wind power, plain bearings, main bearings, pre-loading, run-out Figure 1: Schematic of the FlexPad concept with its key design parameters [5] Wind-Tribology 28 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 side (GS) of the shaft. The RS and GS evaluation points are depicted in Figure 2 (right). They represent the centre points of their respective bearing halves. Furthermore, an evaluation of the hydrodynamic performance of the bearing was conducted for each simulation. 3 Results and discussion In the following chapter the simulation results will be presented and discussed. Focus is on radial run-out, its reduction through preload and the consequences for the hydrodynamic performance of the bearing. 3.1 Shaft run-out As a reduction of shaft run-out is the main motivation for clearance reduction, firstly the bearing’s run-out performance is evaluated for various amounts of positive clearance. In this first stage the bearings performance was investigated for clearances as low as 1 ‰ relative to DOI 10.24053/ TuS-2024-0026 2 Method The FlexPad design used in this contribution was already presented by Rolink et al. [5]. The bearing model is shown in Figure 2 (left). The main design parameters are shown in Table 1. In earlier works nominal clearances between 0.56 ‰ and 2.3 ‰ were investigated for the FlexPad concept [5, 16, 17]. In this work the clearance range is extended to also include preloading. The analysis of the effects of preloading is done via multi-body (MB), elasto-hydrodynamic (EHD) simulations. The model creation was conducted using the toolchain presented in [5]. The simulation setup is identical to [5, 16, 17]. The simulations were performed using the software FIRST. The performance of the bearing was analysed for the static operating conditions described in Table 2. For each bearing simulation the run-out was evaluated. The measured run-out describes the deflection of the shaft respective to its initial position. The run-out was evaluated for the rotor side (RS) and for the generator RS GS Figure 2: FlexPad MB-EHD-model (left), FlexPad shaft with highlighted RS and GS evaluation point for run-out [°] D o [mm] D i [mm] No. Pads [-] Span width [mm] x groove [mm] b groove [mm] t groove [mm] s arm [mm] s pad [mm] 46.7 473.7 256.7 12 249.1 69.7 8.1 9.2 15.2 20 Table 1: Reference design parameters of the investigated bearing rotor speed [rpm] F [kN] F [kN] F [kN] M [kNm] M [kNm] [‰] 28 29 26 2.6 -21 22 0 - 3.5 Table 2: Operating conditions during the EHD simulations Wind-Tribology 29 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 the median diameter. The radial shaft run-out is shown in Figure 3. As expected the shaft run-out increases linearly with the bearing’s clearance. Greater clearance allows for greater shaft movement within the clearance before a hydrodynamically carrying lubrication film is formed. RS radial run-out is smaller than GS run-out. As the run-out decreases linearly with the initial clearance the lowest radial run-out (RS: 0.29 mm and GS: 0.46 mm) is achieved for an initial clearance of 1 ‰. The different scaling of the RS and GS radial run-out for increasing clearance stems from the conical design and the applied load conditions. For the investigated load condition with positive thrust forces the shaft experiences axial movement towards the GS. Due to the conical shape of bearing and shaft this axial movement results in a reduction of clearance on the RS and an increase in clearance on the GS. The increased clearance results in a greater run-out for the GS. Relative operational clearance gained for the GS is roughly equal to the initial clearance of the bearing. This is to be expected, as the shaft movement towards the GS is limited by the RS bearing halve. As positive axial loads are applied the shaft bridges the initial clearance on the RS, thus increasing the GS clearance by the same amount. Similar behaviour was also observed by Rolink for similar designs [18]. The clearance increase for the GS is slightly higher than the initial clearance, due to the flexibility of the bearing, which allows for additional axial shaft movement towards the GS. Preloading, which was achieved via interference fit is also possible for the FlexPad bearing. The negative clearance resulting from the interference fit is equivalent to a preloading of the bearing. The investigated preloads and their respective necessary negative clearance are shown in Figure 4. Preload increases linearly with the simulated negative clearance as the bearing is deformed elastically. For the maximum negative clearance of -1 ‰ the equivalent preload is 460 kN, whereas for -0.1 ‰ the equivalent preload is just 38 kN, which is in the order of magnitude of the applied load case. Preloading the bearing has a positive impact on the shaft run-out. In Figure 5 the radial shaft run-out is depicted for further reduced clearance and increasing preload. Investigated were clearances between the commonly employed clearance of 1 ‰ and -1 ‰, which results in preload due to the previously described interference fit. Reduction below 1 ‰ results in further linear reduced run-out up to zero clearance. Zero clearance in a basic radial plain bearing means that the surfaces of the bushing and the shaft fit together perfectly without contact. Thus, realistically zero clearance cannot be achieved for basic radial plain bearings due to e.g. tolerances and manufacturing inaccuracies. Zero initial clearance is possible for the FlexPad bearing, as during operation the flexible segments DOI 10.24053/ TuS-2024-0026 Figure 3: Radial run-out for RS and GS for varying clearances Figure 4: Corresponding preload for investigated negative clearances Figure 5: Radial run-out for increasing amounts of negative clearance Wind-Tribology 30 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 same is true for the FlexPad concept for which this relationship was already shown by Rolink et al. in [16]. Figure 6 shows the maximum hydrodynamic pressure and the maximum specific pressure for varrying clearances. Maximum specific pressure is defined as the maximum of the normal forces of the most highly loaded segment devided by its surface area. The negative clearances correspond to the preloads depicted in Figure 4. As observed in previous studies, the maximum pressure is reduced for decreasing clearances (see Figure 6). The maximum hydrodynamic pressure decreases from 85 MPa for an initial clearance of 3.5 ‰ to 11 MPa at zero initial clearance. The lowest maximum pressure however, was determined for a negative clearance of -0.12 ‰ or a preload of 48 kN. For larger amounts of preload, the maximum pressure increases. It reaches 89 MPa for a negative clearance of -1 ‰ or a preload of 460 kN. For larger preloads, the bearing becomes hydrodynamically unfeasible. The same applies for the maximum specific pad pressure (see Figure 6). The maximum specific pad pressure decreases for decreasing clearance, from 3.15 MPa at 3.5 ‰ to 1.7 MPa at zero initial clearance. The minimum of 1.66 MPa is reached for a negative clearance of -0.12 ‰ or preload of 48 kN. For increasing preload, the maximum specific pad pressure increases linearly and reaches its maximum at 3.2 MPa for a preload of 460 kN or -1 ‰ negative clearance. To understand the reason for the improved bearing performance under preload, the pressure distribution needs to be evaluated. In Figure 7 the pressure distribution for the RS and the GS of the bearing are shown for a non-preloaded bearing design with 1 ‰ clearance and a preloaded design (-0.12 ‰ negative clearance or 48 kN preload). The applied load conditions are identical. As can be seen, regardless of preload, both bearing designs have hydrodynamic pressure build-up on all RS pads. This is caused by the positive thrust force on the bearing, which presses the RS shaft cone into the bearing, creating sufficiently small lubrication gaps and allowing for hydrodynamic pressure build-up. The maximum specific pad pressure on the RS however, is higher for the preloaded design. It reaches 1.2 MPa for the non-preloaded bearing and 1.5 MPa for the preloaded bearing. However, bearing wide maximum pressures and maximum specific pad pressures are reached at the GS side of the bearing. This occurs regardless of clearance and preload and is the result of the overall bearing design and load conditions. The pressure distribution on the GS is highly influenced by the amount of clearance. For the- DOI 10.24053/ TuS-2024-0026 bent so that the surface is parallel to the counterpart, which allows for desired hydrodynamic load bearing of the FlexPad bearing (see chapter 3.2). For preload the run-out is even further reduced. The shaft run-out decreases asymptotically towards 0.024 mm for the RS and 0.043 mm for the GS at -1 ‰ relative clearance. Thus, large preload decreases the run-out by 92 % for the RS and 91 % for the GS compared to a bearing with 1 ‰ clearance. For larger negative clearance than -1 ‰, the bearing ceases to function as the preloads become too large (see chapter 3.2). Radial run-out decreases asymptotically towards its minimum at -1 ‰ negative clearance or 460 kN preload. Due to fast asymptotical decrease for low amounts of preload, comparatively large amounts of run-out reduction are already possible for small preloads. 72 % of radial-runout reduction for the GS can already be achieved with -0.2 ‰ clearance (or 79 kN preload). The positive effect preloading has on the run-out of the FlexPad bearing is in its cause identical to preloading for tapered rolling bearings. Through preloading the flexible arms of the design are all symmetrically strained. As forces are applied to the shaft, inducing movement, the strain on the arms increases for some and decreases for others. If the shaft would be moved downwards, the downward forces exerted by the upper arms would decrease and the upward forces exerted by the lower arms would increase. The overall bearing therefore generates more upwardly directed forces per increment of downward movement as it would have without the preloaded condition. 3.2 Hydrodynamic performance Naturally the clearance has a significant impact on the hydrodynamic performance of the bearing. Depending on the bearing and on bearing type and design it was shown, that a reduced clearance results in a reduction of the maximum hydrodynamic pressures [19-21]. The Figure 6: Maximum hydrodynamic pressure and maximum specific pad pressure for varying clearances non preloaded bearing design only three pads experience significant hydrodynamic pressure build-up. These three pads subsequently need to carry nearly all the applied loads for the GS. For the preloaded bearing design, seven pads experience significant hydrodynamic pressure build-up. As the GS loads are spread more evenly this results in lower maximum pressures. The poor pressure distribution for non-preloaded bearing designs with clearance has two causes. The first cause is general in nature. An increase in clearance leads to a reduction of pressured area and an increase in maximum pressures. The second cause is specific to the FlexPad concept. Due to the axial shaft movement and subsequent clearance increase described in chapter 3.1 one bearing halve of the conical bearing design always has to operate effectively with double the amount of initial clearance. This is only further exacerbated by the flexible nature of the FlexPad concept, which allows for further axial movement and therefore clearance increase on the GS for the given load case. This phenomenon also explains the improved hydrodynamic behaviour of slightly preloaded bearing designs compared to a design with initial clearance of zero and no preload. The bearing’s flexibility allows for axial shaft movement even for a zero clearance designs. This axial movement leads to an effective clearance increase on the GS of the bearing, which in turn worsens the pressure distribution for this bearing halve. Through preloading the bearing becomes stiffer. This limits the axial shaft movement and maintains minimal amounts of clearance for both bearing halves. This results in an optimum of pressure distribution and maximum pressure. For the investigated design and load conditions this optimum is at 48 kN preload or -0.12 ‰ negative clearance. Although a further increase in preload would further stiffen the bearing and limit its axial movement, this would not lead to a further improvement of maximal pressures since the added loads that need to be hydrodynamically carried by the bearing - stemming from the preload - negate the positive effect regarding the pressure distribution. 3.3 Challenges As the presented study was conducted simulative, no design for a preloading mechanism exist at this stage. Initial preloading of the FlexPad bearing could be achieved via procedures typical for rolling bearings i.e. springs or nuts etc. Furthermore, it could be achieved via an intentional interference fit between conical shaft and bearing. This would require very accurate manufacturing of shaft and bearing to achieve the desired optimal performance. Alternatively, the thermal expansion of shaft and bearing during operations could be used. Differing thermal expansion coefficients or temperature can facilitate a clearance reduction for plain bearings under operating conditions [14, 15]. Therefore, operational preload via thermal expansion could be explored for the FlexPad bearing in future studies. As the FlexPad bearing is envisioned as a WT main bearing the whole bearing life span and its respective operating conditions need to be considered. Most of the operating life of the bearing would be within rated ope- Wind-Tribology 31 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0026 RS GS Ψ=1 ‰, p max =25 MPa Ψ=-0.12 ‰, p max =10 MPa Preload No Preload R G Ψ Figure 7: Hydrodynamic pressure distribution for a non-preloaded (left) with clearance of 1 ‰ and preloaded bearing (right, -0.12 ‰ corresponds to 48 kN preload) Wind-Tribology 32 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 ing reaches its lowest maximum pressure of 10.4 MPa and lowest maximum specific pressure of 1.7 MPa. In total it was shown, that the FlexPad concept can operate hydrodynamically for zero initial clearance and preload. Further studies need to investigate the optimal amount of preload for the whole range of operation conditions for the FlexPad bearing and the best method to apply preload. Furthermore, the feasibility of lubrication film build-up under preloaded conditions needs to be experimentally validated. References [1] EUROPEAN PARLIAMENT: Amendments adopted by the European Parliament on the Renewable Energy Directive (in Kraft getr. am 2022) (2022) [2] 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 [3] ABSCHLUSSBERICHT: Thermisch gespritzte Gleitlagerbeschichtungen für Hauptlager von Windenergieanlagen (WEA) - WEA Triebstrang und Oberflächentechnik, eng.: “Final Report, WEA-GLiTS”: Förderkennzeichen: 03EK3036A [4] SCHRÖDER, Tim Niklas: Konisches Gleitlager für die Rotorlagerung einer Windenergieanlage, eng: “Conical Sliding Bearing for the Rotor Main Bearing of a Wind Turbine” (2021) [5] 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 [6] 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; JACOBS, Georg; BLOCK- MANS, 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-ofthe-art technologies and future development trends. In: Wind Energy Science 7 (2022), Nr. 1, S. 387-411 [7] NEJAD, Amir R.; TORSVIK, Jone: Drivetrains on floating offshore wind turbines: lessons learned over the last 10 years. In: Forschung im Ingenieurwesen 85 (2021), Nr. 2, S. 335-343 [8] EVOLUTION ONLINE: New challenges for rotor bearings in the 8-MW offshore category | Evolution. URL https: / / evolution.skf.com/ new-challenges-for-rotor-bear ings-in-the-8-mw-offshore-category/ . - Aktualisierungsdatum: 2020-04-01 - Überprüfungsdatum 2023-04-26 [9] SKF. URL https: / / www.skf.com/ dk/ news-and-events/ news/ 2019/ 2019-10-08-wind-turbine-main-shaft-bearing -design-considerations. - Aktualisierungsdatum: 2023- 04-27 - Überprüfungsdatum 2023-05-03 DOI 10.24053/ TuS-2024-0026 ration of the WT. The applied loads and rotational speed would therefore be in the range of the investigated load condition. However, turbine start up and extreme loads also need to be considered in the future. The applied preload would further increase the breakaway toque during turbine start-up. This could hamper the turbines ability to idle or start at low wind speeds. Furthermore, the increased load at the start of the turbine could increase time spent in the mixed friction regime before full hydrodynamic load bearing is reached and therefore increase wear. On the other hand, transition speed according to Vogelpohl would theoretically be lowered through the near zero clearance produced by the preload [10, 22]. The general introduction of lubricant within the zero or near zero clearance is as of yet untested through experiments or simulations and needs to be further investigated. 4 Conclusion The FlexPad concept is a promising plain bearing design for future WT main bearings. However, large scale WTs mostly favour highly integrated drivetrain concepts which necessitate strict limitations on maximum main shaft run-out. Plain bearings usually operate with a clearance, which allows free shaft movement within its limits. This creates a conflict between the requirements for main bearings of integrated WTs and the design possibilities with common plain bearings. For rolling bearings, preloading is a common practice to increase stiffness and limit shaft run-out. In the course of this work the effects of clearance reduction and preload were investigated for the FlexPad concept under static operational conditions. Preload was realised via an interference fit resulting in a negative clearance between shaft and bearing. The results of this study show that the stiffness of the FlexPad bearing can indeed be increased through preload while maintaining its hydrodynamic load bearing capability. This also leads to a reduction in shaft run-out for the investigated load conditions. GS run-out reductions of up to 91 % relative to the reference with 1 ‰ relative clearance could be achieved via preload. Too large preloads lead to a breakdown of the hydrodynamic capability of the bearing. It was further discovered, that the bearings hydrodynamic performance could be increased via preloading. This is due to the flexible nature of the FlexPad and its conical design. Preload reduces the clearance increase for the GS through axial shaft movement. This allows for a greater load carrying area on the GS and reduces the overall maximum pressure and maximum specific pressure. For the presented bearing design and load condition an optimal amount of preload was identified with regards to the hydrodynamic performance. The bearing shows optimal hydrodynamic performance for a preload of 48 kN which corresponds to -0.12 ‰ negative clearance. For this design the bear- [10] LANG STEINHILPER: Gleitlager, 1978 [11] WITTEL, Herbert; MUHS, Dieter; JANNASCH, Dieter; VOßIEK, Joachim: Roloff/ Matek Maschinenelemente: Normung, Berechnung, Gestaltung. 22., überarbeitete und erweiterte Auflage. Wiesbaden: Springer Vieweg, 2015 [12] JACOBS, Georg (Hrsg.): Maschinengestaltung. Ausgabe 10/ 2016. Aachen: Mainz, 2016 [13] DIN 31652-3. Januar 2017. Gleitlager - Hydrodynamische Radial-Gleitlager im stationären Betrieb - Teil 3: Betriebsrichtwerte für die Berechnung von Kreiszylinderlagern [14] GROTE, Karl-Heinrich (Hrsg.): Dubbel: Taschenbuch für den Maschinenbau, mit …Tabellen. 23., neu bearb. u. erw. Aufl. Berlin, Heidelberg: Springer, 2011 [15] NIEMANN, Gustav; WINTER, Hans; HÖHN, Bernd- Robert; STAHL, Karsten: Maschinenelemente 1: Konstruktion und Berechnung von Verbindungen, Lagern, Wellen. 5., vollständig überarbeitete Auflage. Berlin, Heidelberg: Springer Vieweg, 2019 (Springer eBooks Computer Science and Engineering) [16] 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 [17] 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 [18] ROLINK, Amadeus; SCHRÖDER, Tim; JACOBS, Georg; BOSSE, Dennis; HÖLZL, Johannes; BERG- MANN, 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 [19] HAGEMANN, Thomas; DING, Huanhuan; RADTKE, Esther; SCHWARZE, Hubert: Operating Behavior of Sliding Planet Gear Bearings for Wind Turbine Gearbox Applications—Part I: Basic Relations. In: Lubricants 9 (2021), Nr. 10, S. 97 [20] KUZNETSOV, Evgeny; GLAVATSKIH, Sergei; FIL- LON, Michel: THD analysis of compliant journal bearings considering liner deformation. In: Tribology International 44 (2011), Nr. 12, S. 1629-1641 [21] SCHILLING, Gregor; LIEBICH, Robert: The Influence of Bearing Clearance on the Load Capacity of Gas Polymer Bearings. In: Applied Sciences 13 (2023), Nr. 7, S. 4555 [22] G. VOGELPOHL: Geringste zulässige Schmierschichtdicke und Übergangsdrehzahl (1962) Initial publication: 71/ 1 Wind-Tribology 33 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 DOI 10.24053/ TuS-2024-0026 TAE-Colloquium Tribology 2024 34 Tribologie + Schmierungstechnik · volume 71 · eOnly special issue 1/ 2024 BOOK RECOMMENDATION expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. 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ISSN 0724-3472 Science and Research www.expertverlag.de Philipp Harrer, Dmitrii Svetov, Patrick Eisner, Maximilian Lackner The Effects of Applying the Tribological Compound TZ NIOD - reversing wear Jörg Loos, Wolfram Kruhöffer, Daniel Merk, Toni Blaß, Jörg Franke Reibungsbedingte WEC-Bildung bei hohen Lasten Ivan Grozev, Sagar Dalal, Nazlim Bagcivan, Serhan Bastuerk, Christian Lueffe, Thomas Stahl Simplified tribological approach for predesign of wind turbine bearing cases, combined with model test investigation Jan Euler, Georg Jacobs, Timm Jakobs, Julian Röder Feasibility study on preloaded flexible conical plain bearings as wind turbine main bearings