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Tribologie und Schmierungstechnik EDITOR IN CHIEF MANFRED JUNGK 1 _ 24 VOLUME 71 Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie An Official Journal of Österreichische Tribologische Gesellschaft An Official Journal of Swiss Tribology Issue 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. When no source is indicated, all rights to pictures are reserved by the author or the editorial office. No third-party claims can be made unless otherwise agreed upon. The editorial office retains the right to edit and shorten articles. Trade names and commercial names mentioned in this journal may not be readily used by everyone, as they are often registered and protected trademarks. The journal, including all articles and pictures, is protected by copyright law. Excluding legally permitted cases, further use of the content without the publisher’s consent is punishable by law. This applies especially to copying, translating, creating microfilms, and using and processing the content in electronic systems. All information in this journal has been compiled with great care. However, mistakes cannot be ruled out entirely. Therefore, neither the publisher nor the authors assume liability for the correctness of the content or any mistakes and their consequences. 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ISSN 0724-3472 ISBN 978-3-381-11581-5 Impressum Tribologie und Schmierungstechnik Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie | An Official Journal of Österreichische Tribologische Gesellschaft | An Official Journal of Swiss Tribology Editorial 1 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0001 Over the years we have published more and more articles in English and have decided from volume 71 onwards to publish the entire journal in the globally most used language. Besides the necessity to gain a higher impact factor and thus more citations for our authors it was also the German Tribological Societies Technical Advisory Board that was in favour of that decision. Publications purely in the German language become rarer and rarer in the scientific and industrial world. Offering tribological research and technology in the centre of Europe to a broader audience will benefit readers and authors. And yes, there is always a downside to a decision. Reading a journal in your mother tongue is more comfortable compared to a foreign language. Personally, I remember very well back in my university time during the third semester when my professor for Physical Chemistry chose a book in English to accompany his class. Initially he was not very popular among us students, in the end I conducted my doctorate studies at his institute and read more English books and articles. My next encounter with English in the professional world was back in 1988 as I started to work for a multinational company at a smaller site in Germany. Back then to me it was great to communicate via a main frame computer to share calendars and write emails - the younger ones amongst us probably do not know why I’m mentioning that. My German colleagues were questioning me why I would write the emails in English. The easy answer is I wanted colleagues around the globe to be able to read them. Since having taken over the editorship with Volume 66 in 2019 we introduced the possibility to read the journal online besides the print version, added the peer review process as well as open access option and started publishing two extra E-only issues per year. Thus, switching to English is the next logical step toward a leading tribology journal. We will keep the name “Tribologie und Schmierungstechnik” and add the subtitle “Tribology - Lubrication Friction Wear”. As we are in the transition phase, we will also change the sections patents and standards to English. Tribology relevant reviews of German Patent publications and DIN’s Beuth Verlag will be complimented with publications of the European Patent Office and ISO. At the end of this editorial, I would like to point out that in 2020 expert verlag started the quarterly journal “Schmierstoff und Schmierung”, which is a publication dealing with application engineering in German language only with mother tongue comfort. Enjoy studying this issue and remember Tribology is everywhere. Your editor-in-chief Manfred Jungk “Tribologie und Schmierungstechnik” goes International! Events 2 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 Events Date Place Event ► 19.05. - 23.05.24 Minneapolis, 78 th STLE Annual Meeting & Exhibition Minnesota (USA) ► 11.06. - 14.06.24 Lyngby, Denmark Nordic Tribology Symposium 2024 (NORDTRIB 2024) ► 18.06. - 20.06.24 San Sebastian, Spain Lubricants, Tribology and Condition Monitoring (LUBMAT 2024) ► 22.07. - 23.07.24 Erlangen, Germany 7 th Young Tribological Researcher Symposium 2024 ► 02.09. - 04.09.24 Lyon, France 49 th Leeds Lyon Symposioum on Tribology ► 17.09. - 19.09.24 Messe Düsseldorf, Lubricant Expo Europe Germany ► 23.09. - 25.09.24 Göttingen, Germany 65 th German Tribology Conference 2024 TuS PLUS: Tribologie und Schmierungstechnik jetzt mit noch mehr Fachinformation online Ab diesem Jahr erscheinen von der „Tribologie und Schmierungstechnik“ zwei zusätzliche Ausgaben jährlich. 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Abo-Service: Tel: +49 (0)7071 97 97 10 eMail: abo@narr.de Contents 3 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 5 Zita Tappeiner, Achill Holzer, Katharina Schmitz Experimental development and validation of tribological run-in strategies to reduce friction in hydraulic applications 14 Jan Euler, Georg Jacobs, Timm Jakobs, Julian Röder Feasibility study on preloaded flexible conical plain bearings as wind turbine main bearings 22 Magnus Schadomsky, Johann Rauhaus, Lars Blumenthal, Detmar Zimmer, Balázs Magyar Investigation of the friction and deformation behaviour of high-speed brakes 35 Linda Becker, Peter Tenberge Local considerations and experimental results on the contact behavior of crossed helical gears with general flank geometries 1 Editorial “Tribologie und Schmierungstechnik” goes International! 2 Events Science and Research 43 News Gesellschaft für Tribologie Columns Preface Tribologie und Schmierungstechnik Tribology - Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie An Official Journal of Österreichische Tribologische Gesellschaft An Official Journal of Swiss Tribology Volume 71, Issue 1 May 2024 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). IHR ONLINE-ABONNEMENT DER TuS Ab dem Jahrgang 2019 können Sie die aktuellen Hefte der Tribologie und Schmierungstechnik im Online-Abonnement beziehen. Die Hefte der vergangenen Jahrgänge werden kontinuierlich integriert. Unsere eLibrary bietet Ihnen einen qualitativ hochwertigen und benutzerfreundlichen Zugang zum digitalen Buch- und Zeitschriftenprogramm der Verlage expert, Narr Francke Attempto und UVK. Nutzen Sie mit uns die Chancen der Digitalisierung: https: / / elibrary.narr.digital/ journal/ tus Der Online-Zugang ist in Kombination mit dem Print-Abo oder als e-only-Abo erhältlich. Abo-Service: Tel: +49 (0)7071 97 97 10 Fax: +49 (0)7071 97 97 11 eMail: abo@narr.de Anzeige 4 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 Eine Zeitschrift des Verband Schmierstoff-Industrie e. V. SCHMIERSTOFF SCHMIERUNG www.sus.expert Hier können Sie die Zeitschrift kostenlos abonnieren. E R S C H E I N T V I E R M A L I M J A H R Introduction Axial piston pumps are hydraulic displacement units that are widely used in both industrial and mobile applications due to their high power density and compact design / Mat14/ . In an axial piston pump, there are many tribological contacts. This paper concentrates on the tribological contact between cylinder block and valve plate. The contact has several similarities with axial thrust bearings but due to the valve geometries there are also other challenges to be dealt with. During operation, the cylinder block is preloaded against the valve plate to reduce leakage. The pressurization of the pistons and load shocks from the engine generate additional axial loads. A widely used type of axial piston pump is the swash plate machine. A detailed view of the components of this sub-type is shown in Figure 1. There is a combination of hydrodynamic and hydrostatic effects in the contact between the cylinder block and the valve plate. The valve plate is hydrostatically balanced by pressurized oil, creating a thin oil film between the surfaces. However, the resulting axial load is dependent on the piston position, which leads to a periodic pulsation of the gap height. In addition, a hydrodynamic lubricating film builds up during operation. Due to changes in load and rotational speeds, the balancing of the contact will usually not be ideal in all states of operation, resulting in the hydrodynamic lubricating film failing at low speeds and mixed friction occurring. / Sei08/ . In this pa- Science and Research 5 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0002 Experimental development and validation of tribological run-in strategies to reduce friction in hydraulic applications Zita Tappeiner, Achill Holzer, Katharina Schmitz * submitted: 20.09.2023 accepted: 22.01.2024 (peer-review) Presented at the GfT Conference 2023 Das Einlaufen eines tribologischen Kontakts hängt von der aufgebrachten Normallast, der Gleitgeschwindigkeit und der Einlaufdauer ab. Um Axialkolbenpumpen effektiv einzulaufen, suchen die Hersteller nach optimalen Einlaufparametern. In diesem Paper wird eine Methodik entwickelt, um geeignete Einlaufparameter zu bestimmen. Die Parameter wurden an einem Scheibe-Scheibe-Tribometer validiert und bestätigt. Ziel der Untersuchung ist die Minderung von Reibung und Verschleiß im späteren Pumpenbetrieb. Schlüsselwörter Axialkolbenpumpe, Hydraulik, Einlaufen, Stribeck- Kurve, Tribometer, Scheibenkontakt A particularly demanding contact in axial piston pumps is the one between cylinder block and valve plate. The tribological behavior of the contact can be changed by a run-in process. Publications on fast and efficient run-in are rare as this knowledge is often considered confidential. For this paper, tests have been carried out on a disc-on-disc tribometer to examine the run-in behavior of a material pairing and to identify suitable parameters for run-in. A methodology has been developed which can be used to find the optimal normal load and rotational speed for the runin process to shorten the run-in time significantly. Keywords Axial Piston Pump, Disc-On-Disc, Hydraulics, Run- In, Stribeck Curve, Tribometer Kurzfassung Abstract * Zita Tappeiner, M.Sc. Orcid-ID: https: / / orcid.org/ 0009-0008-7442-5809 Achill Holzer, M.Sc. Orcid-ID: https: / / orcid.org/ 0000-0003-1190-1819 Univ.-Prof. Dr.-Ing. Katharina Schmitz Orcid-ID: https: / / orcid.org/ 0000-0002-1454-8267 Institut für fluidtechnische Antriebe und Systeme der RWTH Aachen University Campus-Boulevard 30, 52074 Aachen on the influence of the normal force and the rotational speed occurring during operation. In addition, it is to be determined how the friction and wear during operation can be reduced using different run-in parameters. In tribology, the term “run-in” is used to describe the processes that occur during the initial encounter of two friction partners until a steady-state friction condition is reached. These processes are accompanied by a change in the coefficient of friction and/ or the wear rate as a function of the run-in time, the number of friction cycles or the friction distance. Run-in is used specifically for surface conditioning in tribosystems / Bla91/ . Surface conditioning is achieved by changing the mechanical, chemical and physical properties as well as the contact geometry of the friction partners. This process takes place on the surface and in the microstructure near the surface / Fes14/ . The three most important characteristics of friction transitions are the change in the coefficient of friction, the time required for a tribosystem to reach a steady state (or other distinctive state such as galling), and the characteristics during short-term fluctuations in the frictional load. Blau identified common shapes of run-in curves / Bla08/ . Figure 2 shows the variation of the frictional force typically occurring in pure metal contacts. The shape shown is called form (f) by Blau. The decrease in frictional load occurs due to changes in the near-surface layers and the smoothing of the surface. Overall, four processes can be identified that constitute run-in: Material transfer, film Science and Research 6 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0002 per the tribological behavior of the contact between cylinder block and valve plate is examined. Its focus lies Figure 1: Axial piston pump in swash plate design Figure 2: Friction behavior during the run-in of metal contacts according to / Bla08/ Figure 3: Combined effect of load and speed on the surface roughness variation / Akb13/ formation or removal, deposition, and cyclic surface deterioration. The curve shape shown cannot be clearly assigned to mechanisms, and similar curve shapes can be caused by different boundary processes / Bla08/ . The parameters selected during run-in affect the performance of the tribological system in operation. Accordingly, the operating conditions during the run-in phase should be carefully selected. Optimizing the run-in process can lead to an extension of a tribo system’s service life and stable operation. Design, surface mechanics, chemistry and materials play important roles in optimization / Kho21/ . Akbarzadeh and Khonsari determined the operating conditions that result in the minimum value of the surface roughness R a at the end of the run-in period. Figure 3 shows the effect of the combination of load and speed on the change in surface roughness / Akb13/ . It is noticeable that the operating conditions for the maximum change in surface roughness are very close to operating conditions that cause a significantly smaller change in surface roughness. In addition to load and speed, the initial surface roughness and the initial friction coefficient, which in turn depends on load and speed, are also decisive / Kho21/ . Methods Tribometer Model Test Rig Figure 4 shows the setup of the disc-on-disc tribometer test rig (tribometer) also known as Siebel-Kehl tribometer / Mur10/ , used for the experimental part of this work. The two ring-shaped test specimens “stator” and “rotor”, which are pressed onto each other with a defined load by the hydraulic contact cylinder, are the central elements of the tribometer. A load cell between the cylinder and the moment support is used to measure the contact pressure. A hydraulic motor generates the necessary relative movement of the rotor. The frictional torque between the rotor and stator is determined on the stationary, upper disc via a force sensor connected to a lever. The test specimens are selected according to the material pairing in axial piston pumps. The stator is made of heat-treated steel (1.7225) and the rotor of the special brass (2.0550). A mineral oil (Renolin B 15 VG 46) containing zinc and ash of viscosity class 46 was used for the tests / Fuc24/ . The fluid contains a zinc dithiophosphate as extreme pressure and antiwear additive. All test runs were repeated with three specimens of the same material combination. After each test run, the surface of the discs was renewed using wet sandpaper with a final grit of 1200. The test bench can reach rotational speeds of up to 1600 min -1 . With an average diameter of the contact surface of 27 mm, an average speed of 4.5 m/ s is achieved. Loads of up to 3500 N were used, inducing surface pressures of up to 4 MPa. In a load analysis on an axial piston pump, Wegner determined the surface pressure at the control surface to be 4 MPa. / Weg21/ Thus, the parameters used are within the range of values that occur in the hydraulic application. Experimental approach In a first approach, the tribological contact was examined under constant conditions. In a test run lasting one hour, temperature and frictional torque in the contact between the discs were measured. The test parameters are listed in Table 1. Science and Research 7 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0002 Figure 4: Tribometer test rig Oil temperature 40 °C Test duration 60 min Rotational speed 1600 rpm Normal load 1000 N Table 1: Test parameters for constant tests Subsequently, three tests were carried out in which either the normal load or the rotational speed were changed step by step. This procedure was chosen in order to be able to compare the measured friction coefficients of At 1600 min -1 , the normal load was adjusted to 1000 N. Finally, the Stribeck curve was measured following a speed ramp (duration: t Stribeck . The following values for the run-in time t run-in were tested: 0 min, 1 min, 3 min, 5 min, 10 min and 30 min. These tests were carried out for the normal loads F N,run-in 1000 N, 2000 N and 2500 N. Results Figure 6 shows the results for 1000 N normal load (corresponding to a surface pressure of about 1.2 MPa). The coefficient of friction drops at the beginning of the measurement and remains constant. The coefficient of friction of pairing no. 3 is slightly lower overall and shows hardly any fluctuations over the entire period. The mean value for the coefficient of friction is 0.0322 at the beginning. Within one hour at constant test conditions, the mean coefficient of friction drops to 0.0300. This corresponds to a 7 % decrease in the coefficient of friction. A comparison of the different normal forces shows that the measured coefficient of friction decreases as the contact pressure increases. This contrasts with the findings of Popov / Pop15/ . In Bollók’s investigations with a steelbronze pairing, however, the same dependence can be Science and Research 8 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0002 two stages with the same test parameters. In these three tests, the influence of normal load and speed was investigated iteratively. An overview of these tests can be found in Table 2. Test parameters Test 1 Test 2 Test 3 Oil temperature 40 °C 40 °C 40 °C Holding time per step 5 min 5 min 5 min Rotational speed 1600 rpm variable 250 rpm Normal load variable 2000 N variable Step-size 250 N 125 rpm 250 N Initial value 250 N 1500 rpm 250 N Maximum value 3500 N 1500 rpm 3500 N Table 2: Test parameters for variable tests In a final series of tests, specific run-in conditions were compared. To compare the quality of the run-in process, the same Stribeck curve was repeated each time. The test load and speed procedure are shown in Figure 5. In the beginning of the measurement, the speed was increased to 250 min -1 with a defined slope. After that, the normal load was set according to the run-in test. The run-in then took place at constant conditions for the time t run-in . Hereafter, the speed was increased to 1600 min -1 . Figure 6: Characterization of the tribological contact Figure 5: Test conditions for Stribeck tests observed. As the normal force increases, the coefficient of friction decreases / Bol06/ . Figure 7 shows the change in the disc surface due to the run-in process. Abrasive running marks occur in the contact area with the other disc. In some cases, even deep ridges are formed, which are either due to roughness peaks of the counter body or particle contamination in the fluid. No adhesion is recognizable in the surface after run-in. A detailed examination of the surface roughness was not part of this work but could provide further information on the run-in condition of the components. The measurement results in Figure 8 show that the pairings tolerate a normal load of 3500 N, as no fretting occurs in pairings 1 and 3 under these test conditions. In pairing 2, on the other hand, a sharp increase in the coefficient of friction and, analogously, in the temperature already occurs at 1750 N. The reason for this is unclear, since in a preliminary study pairing 2 withstood a normal load up to 3000 N without an increase in the coefficient of friction. Thus, pairing 2 is considered as an outlier here. The other two pairings exhibit a different, very similar behavior to each other. In both pairings, the coefficient of friction decreases as the normal load increases. At the beginning of each new load level, the coefficient of friction increases slightly and then decreases again shortly afterwards. This is an indication that run-in takes place at each step. After the first half of the gradational progress has been run through and the load decreases again, these increases no longer occur. Overall, the values of the coefficient of friction as well as the temperature on the right-hand side of the gradational progress are lower than on the left-hand side. In order to obtain a quantitative measure for the run-in, a run-in quotient μ after / μ before is introduced. The run-in quotient is formed according to the illustration in Figure 8. For each load level, the friction coefficient is averaged. Then, for each load step, the mean value μ after from the right side of the stair is divided by the corresponding mean value μ before from the left side. Thus, a quotient of the improvement of the friction coefficient can be assigned to each value of the normal load. The graphical representation shows that the coefficient of friction decreases for almost all load levels over the duration of the test. At this point, it should be noted that the time difference between the left and right sides of the gradational progress is not constant for the individual load steps. Due to the stepwise test methodology, the time difference and thus the run-in time is significantly larger for the lower values of F N . Nevertheless, an improvement in the coefficient of friction can be observed, especially for the Science and Research 9 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0002 Figure 8: Influence of the normal force on the run-in behavior Figure 7: Change in the rotor surface due to run-in difficult to evaluate. Nevertheless, it is noticeable that despite the fretting, there was an improvement in the coefficient of friction. At this point, it is important to note that the improvement is only evaluated against the same normal load. The following experimental investigations are evaluated analogously to the procedure in Figure 8. The method can be applied not only to the tests with varying normal loads, but also with varying speeds. The aim here is to find the optimal parameters for the run-in of the contact. Since the run-in process depends on a combination of the parameters speed and load, this investigation must be carried out iteratively. The plots of the run-in quotients of experiment 2 and 3 in Figure 9 show clear minima at 250 rpm (0,71 m/ s) and between 2500 and 3000 N respectively. To verify an Science and Research 10 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0002 higher normal loads. Since the quotient is formed from μ after / μ before , the greater the change, the smaller the value of the quotient. Thus, a smaller value of the run-in quotient is an indication of a greater improvement in the coefficient of friction. Values higher than 1 indicate an increase in the coefficient of friction and thus a worsening of the friction behavior. This is the case with the lower load values of pairing 3. Why a deterioration of the friction condition occurred in this case, cannot be explained at this point. For other normal loads and pairings, a general decrease in the coefficient of friction can be observed. It is interesting to note that the progressions of the curves of pairings 1 and 2 are very similar. Both have a global minimum at about 2500 N. Apparently, run-in under a normal load from this range of values leads to a stronger improvement of the frictional properties than a run-in at values around 1000 N. Since fretting occurred in pairing 2, the values of the run-in quotients are more Figure 10: Validation of run-in parameters Figure 9: Run-in quotients of Test 2 (left) and Test 3 (right) improvement in the run-in, validation was carried out with Stribeck tests. The results of the validation are shown in Figure 10 by the normalized friction coefficient. For better comparison the friction coefficient is normalized with the maximum friction coefficient μ ref . Evidently, the parameters for major changes in the coefficient of friction determined by means of the run-in quotient do indeed lead to better run-in results. In particular, the extreme reduction of the run-in time from run-in at unfavorable conditions (1000 N) to run-in at more favorable conditions (2000 N and 2500 N) poses many economic and ecological advantages. The final validation is shown in Figure 11. The Stribeck test was carried out at different normal loads after a runin time of three minutes in each case. A further increase in the normal force does not lead to an improvement in the run-in process, but to higher coefficients of friction than the running-in at 2000 N or 2500 N. This means that an optimum can indeed be determined by the developed methodology. Discussion A general methodology for identifying suitable run-in parameters for a tribological contact can be derived from the procedure determined in this work empirically. Since a tribological system depends on many factors, such as geometry, material, lubricant and temperature, the runin parameters developed during the experimental investigation cannot simply be transferred to any tribological system. However, the findings developed here can be used to characterize a new tribological system. An illustration of the procedure followed in this investigation is shown in Figure 12. Starting from a variation of the normal load at an initially randomly selected speed (Test 1), a normal load suitable for the run-in was determined. With the determined load, the speed achieving the greatest run-in effect was sought (Test 2). The initial test was then repeated at the new speed (Test 3). Thus, three trials with two transmissions (I and II) took place. Following this procedure, suitable run-in parameters can already be identified. However, it is advisable to supple- Science and Research 11 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0002 Figure 12: Method for creating a run-in map Figure 11: Validation of the influence of the normal load on the friction coefficient and rotational speed can be determined for any tribological contact. This is particularly useful for new material pairings in order to improve the friction coefficient and reduce the run-in time. References / Akb13/ Akbarzadeh, S., Khonsari, M. M. On the optimization of running-in operating conditions in applications involving EHL line contact, Wear, Vol. 303, 1-2, S. 130-137, 2013. / Bau20/ Bauer, G., Niebergall, M. Ölhydraulik - Grundlagen, Bauelemente, Anwendungen, Lehrbuch, 12., neu bearbeitete Auflage, Springer Vieweg, Wiesbaden, Heidelberg, 2020. / Bla91/ Blau, P. J. Running-in: Art or engineering? Journal of Materials Engineering, Vol. 13, Nr. 1, S. 47-53, 1991. / Bla08/ Blau, P. J. Friction science and technology - From concepts to application, Dekker Mechanical Engineering, 2nd ed., CRC Press, Boca Raton, London, 2008. / Bol06/ Bollók, P., Kozma, M. Comparison of surface layers developed during sliding friction of metal pairs, International Colloquium Tribology, 2006. / Fes14/ Feser, T. Untersuchungen zum Einlaufverhalten binärer alpha-Messinglegierungen unter Ölschmierung in Abhängigkeit des Zinkgehaltes, Schriftenreihe des Instituts für Angewandte Materialien, Karlsruher Institut für Technologie, Print on demand, KIT Scientific Publishing, Karlsruhe, 2014. / Fuc24/ RENOLIN B 15 VG 46, https: / / www.fuchs.com / de/ de/ produkt/ product/ 149041-RENOLIN-B-15- VG-46/ (abgerufen am 22.01.2024), 2024. / Kho21/ Khonsari, M. M., Ghatrehsamani, S., Akbarzadeh, S. On the running-in nature of metallic tribo-components: A review, Wear, 474-475, S. 203871, 2021. / Mat14/ Matthies, H. J., Renius, K. T. Einführung in die Ölhydraulik - Für Studium und Praxis, 8. Auflage, Springer Vieweg, Wiesbaden, 2014. / Mur10/ Murrenhoff, H. Umweltverträgliche Tribosysteme - Die Vision einer umweltfreundlichen Werkzeugmaschine, Springer-Verlag Berlin Heidelberg, Berlin, Heidelberg, 2010. / Pau17/ Paulus, A. Reaktionsschichtbildung auf bleifreien Bronze- und Messingwerkstoffen im Kontakt von Zylinder und Steuerscheibe einer Axialkolbenpumpe, 2017. / Pop15/ Popov, V. L. Kontaktmechanik und Reibung - Von der Nanotribologie bis zur Erdbebendynamik, 3., aktualisierte Auflage, Springer Vieweg, Berlin, Heidelberg, 2015. / Sei08/ Seifert, V., Alaze, N. et al. HYDRANO - Leistungssteigerung hydraulischer Verdrängereinheiten durch Nanocomposites - Schlussbericht. Projektlaufzeit: 01.06.2005 bis 30.11.2008, S. 1-236, 2008. / Weg21/ Wegner, S. Experimental and Simulative Investigation of the Cylinder Block/ Valve Plate Contact in Axial Piston Machines, Reihe Fluidtechnik D, Shaker, Aachen, 2021. Science and Research 12 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0002 ment these tests with a further test. This is shown in black in Figure 12 and forms the fourth edge, which is necessary to create a run-in map. In this way, a graphical representation of the inlet quotient could be developed analogous to the change in surface roughness in Figure 3. This map could be supplemented as required in further tests, resulting in an increasingly accurate characterization of the run-in behavior of the tribological system. The method developed here can be applied to discs of different materials using different oils at any oil temperature on the tribometer used in this work. With the aid of the method developed here, it is then possible to determine run-in parameters that can be used for all similar applications. Thus, the often very time-consuming and energy-intensive run-in processes in industry could be significantly optimized. Conclusion In this work, run-in examinations have been carried out to improve run-in behavior of tribological contacts by identifying optimal parameters for normal load and rotational speed. With the identified parameters both the friction coefficient could be decreased and the run-in period shortened. The influence of speed and normal load on the run-in process was analyzed using a gradational progress on a tribometer model test rig. In order to measure the run-in progress, a run-in quotient was introduced, as a measure of the change in the coefficient of friction before and after the run-in process. A comparison of the parameters rotational speed and normal load shows the influence of normal load on the run-in behavior being significantly higher than the influence of rotational speed. The developed methodology was validated in further experiments. Run-in tests at various normal loads were carried out at the optimum run-in speed. These confirmed that run-in at the optimum normal load determined following the methodology leads to a more significant improvement in friction behavior than run-in at other normal loads. A 7 % decrease in the coefficient of friction at the low point of the Stribeck curve was achieved. In addition, with the correct choice of run-in parameters, the duration of run-in could be significantly reduced. Using the optimal parameters, the run-in can be completed within 3 min. With this finding, significant time savings are possible compared to the usual run-in period in practice. Following the scheme of the methodology developed in this work, a suitable set of parameters for normal load \ Gesundheit \ schaft \ Linguisti schaft \ Slawisti \ Sport \ Gesun wissenschaft \ L wissenschaft \ philologie \ Spo Fremdsprachend \ VWL \ Maschi schaften \ Sozi Bauwesen \ Fre Science and Research 13 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikationswiss chaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwiss chaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Tourismus \ VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ Altphilol Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikatio issenschaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Spra issenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Tourismus \ VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ hilologie \ Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwese remdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwissenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Touris VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ Altphilologie \ Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwiss chaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikationswissenschaft \ Linguistik \ Literaturgeschichte \ Anglisti auwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwissenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtsc BUCHTIPP Rüdiger Krethe Handbuch Ölanalysen 1. Auflage 2020, 284 Seiten €[D] 148,00 ISBN 978-3-8169-3499-8 eISBN 978-3-8169-8499-3 expert verlag - Ein Unternehmen der Narr Francke Attempto GmbH + Co. KG Dischingerweg 5 \ 72070 Tübingen \ Germany Tel. +49 (0)7071 97 97 0 \ Fax +49 (0)7071 97 97 11 \ info@narr.de \ www.narr.de Das Buch bietet eine praxisorien琀erte Einführung in das Thema Ölanalysen. Es vermi琀elt das nö琀ge Hintergrundwissen, von der sachgerechten Probenentnahme, den Prüfverfahren bis zum Verstehen der Analysenergebnisse. Hierdurch unterstützt es den Anwender dabei, kostspielige Ausfallzeiten der Maschinen zu verhindern. Rüdiger Krethe ist diplomierter Maschinenbauer und Tribotechniker. Er befasst sich seit mehr als 25 Jahren intensiv mit der Schmierung von Maschinen, angefangen von der Produktauswahl, der innerbetrieblichen Organisa琀on bis hin zur Überwachung von Schmierölen und Hydraulik昀üssigkeiten während des Einsatzes. Science and Research 14 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0003 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 Jocobs Timm Jokobs, 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. Science and Research 15 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0003 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] 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 Science and Research 16 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0003 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 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 Science and Research 17 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0003 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 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- Science and Research 18 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0003 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- Science and Research 19 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0003 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) 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 Science and Research 20 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0003 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) Science and Research 21 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0003 3,000 rpm up to 30,000 rpm. Typical stationary applications include spindle drives or turbo compressors. They are increasingly being used as high-speed direct drives, Science and Research 22 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 Introduction and motivation Brake rotors and static pressure plates, which are pressed together during the braking process, are a constructive component of a large number of common brake designs. According to the classification by S CHLECHT , B. [1], these brakes fall under the category of multi-disc brakes with axially displaceable braking surfaces. Common types include multi-disc solid brakes, permanent magnet brakes and spring-applied brakes, which cover a wide range of applications from industry to e-mobility and materials handling. Miniaturisation and cost reduction are development efforts in drive technology across all industries and thus represent an important aspect in the development of brakes. Particularly due to the increasing spread of electrically controlled drives that can also decelerate drive trains recuperatively, permanent magnet brakes and spring-applied brakes are now mainly in demand as emergency and holding brakes. Although these are necessary, they are rarely used and are therefore subject to increased cost and installation space pressure [2], [3]. The increasing number of mobile applications also promotes the demand for brakes of increasingly high power density [4]. In addition, the field of application of high-speed drives is constantly expanding due to their favourable ratio of material used to output power; they are now available in a wide power range from 0.1 kW to several MW [5]. High-speed drives have a much higher rated speed than the typical industrially used speeds, from well over Investigation of the friction and deformation behaviour of high-speed brakes Magnus Schadomsky, Johann Rauhaus, Lars Blumenthal, Detmar Zimmer, Balázs Magyar* submitted: 18.09.2023 accepted: 19.03.2024 (peer-review) Presented at the GfT Conference 2023 Spring-applied brakes are widely used components in industrial drive systems. They provide a braking torque by friction between a mostly organic friction lining and a metallic counter surface. Increasing with decreasing size, they currently achieve speeds of up to 6,000 rpm, which corresponds to a sliding speed in frictional contact of up to 35 m/ s. At the same time, there is a trend towards high-speed drives, with speeds of 10,000 rpm and above. So far, little is known about the behaviour of the friction value and torque of conventional spring-applied brakes with low-cost organic friction linings under these operating conditions. For this reason, a test rig was developed that allows testing at sliding speeds of up to 120 m/ s with different load inertias. The tests carried out at KAt so far showed that with limited friction work, the conventional spring-applied brake reaches the nominal braking torque at higher sliding speeds. In addition to thermal overload of the friction lining, plastic deformation of the friction bodies can also permanently disrupt the operating behaviour of brakes operated at high sliding speeds. The plastic deformation of the friction discs manifests itself, for example, in a saucer-like shape of the discafs, leading to a reduction in the air gap and causing unwanted changes in the friction conditions. This paper describes the relationship between friction work and friction coefficient in organic linings and the physical mechanism of the deformation process of the friction discs. Based on these possible measures to reduce deformation are explained. Keywords spring-applied brake, brakes, High speed brakes, High speed friction, saucer-shaped deformation, thermal deformation Abstract * Magnus Schadomsky, M.Sc. Johann Rauhaus, M.Sc. Lehrstuhl für Konstruktions- und Antriebstechnik (KAt), Universität Paderborn, 33098 Paderborn Lars Blumethal, M.Sc. Prof. Dr.-Ing. Detmar Zimmer, ehemals Lehrstuhl für Konstruktions- und Antriebstechnik (KAt), Universität Paderborn, 33098 Paderborn Prof. Dr.-Ing. Balázs Magyar Lehrstuhl für Konstruktions- und Antriebstechnik (KAt), Universität Paderborn, 33098 Paderborn which means that the gearbox which translates into high speed can be dispensed with. [5,6] This trend is supported by the development of high-speed asynchronous machines [6,7,8], which in relation to other types of drives are characterised by their simple design and the resulting low-cost production [9]. In the field of brake motors, high-speed drives have not yet become established. On the one hand, this is due to the often non-existent speed requirement. Furthermore, the brakes typically used are not designed for high speeds. In applications where installation space and mass are not an essential criterion, conventional brake motors will continue to be used because of the low-cost standard asynchronous motors. However, fast-running brake motors become technically relevant in applications in which high dynamics play a role and the drive itself is also moved. Furthermore, in geared motors, increasing the transmission ratio in conjunction with a faster-running motor can have a positive effect on the costs of the drive system by saving copper in the motor and brake. Given the outlined application scenarios, the development of brakes with high power density and thus high sliding speeds in frictional contact becomes relevant. This is not only a challenge from a design point of view; due to the scaling of the size with constant power, various physical problems arise. The following types of damage are described in the literature: • Hotspots: Point by point, significantly increased temperatures on the brake disc. A NDERSON , E. et al [10] describe four different types of hotspots and their consequences, e.g. structural transformation. • Hotbands: Ring-shaped areas of significantly increased temperature [11]. • Sinter carry: Increase in the coefficient of friction due to seizure of friction material and the brake disc [12]. • Fading: Collapse of the coefficient of friction due to overload and thus damage to the friction material [13]. • Melting of the friction lining in the case of organic friction materials and transfer to the pressure plate. • as well as lining detachment in the case of lining discs [14]. The aim of this paper is to investigate the potential of the currently used standard spring-applied brakes with their low-cost organic friction linings for applications in higher speed ranges. Furthermore, the deformation of the pressure plates due to high temperatures in frictional contact will be investigated. State of science and technology Spring-applied brakes The spring-applied brake is mainly mounted directly on the B-side of an electric motor. The design corresponds to the disc brake, the braking force is provided by pretensioned compression springs acting on the armature disc (Figure 1). The braking torque is generated by frictional engagement between the friction linings of the rotor, the armature disk and the B end shield of the motor, which is usually protected by a friction plate. The friction system generally consists of organic friction linings in combination with steel or cast counter surfaces [15]. The brake is released electromagnetically by energising the coil. Thus, the brake is linked (closed) in the de-energised state. This design implements the closed-circuit current principle. Figure 1 shows the two possible switching states of a conventionally designed spring-applied brake. The braking torque of a spring-applied brake is calculated according to (1). (1) Legend: T B Nominal braking torque F F Spring force r m Friction radius µ Friction coefficient z Number of friction surfaces = Science and Research 23 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 Figure 1: Basic construction and mode of operation of a spring-applied brake cient of friction collapses. The damage to surface elements is also the mechanism that determines wear. The decisive factor for the thermal load of a friction lining is the friction power, which in turn is influenced by other variables such as friction speed, contact pressure or inertia. It decisively determines the temperature in the friction contact and thus influences the balance between damage and regeneration of the surface elements. [18] It is known that an impermissibly high friction work during braking leads to a collapse of the friction coefficient and subsequently the braking torque. High temperatures caused by excessive friction power can also have a negative effect on the friction coefficient. [16, 18] In general, the thermal load of friction linings is determined by the following variables: - Temperature in frictional contact, - Specific friction work (2), - Specific friction power (3), whereby the respective influence depends on the load constellation. [18] (2) (3) Legend: q Specific friction work J Inertia ω Angular speed A Friction surface of a brake lining q˙ Specific friction power T B Brake torque z Number of friction surfaces Above approx. 350 °C, the organic compounds of the friction linings decompose, which is why other materials are used for application scenarios with very high thermal loads, such as racing clutches. Sintered metal, carbon fibre or ceramic materials can withstand higher temperatures, but have a significant cost disadvantage. [17,20] A LBERS , A. et al [20] shows that there is a need for research into the systematic understanding of the performance limits of organic friction linings and has developed a test concept in which sliding speeds of up to 40 m/ s can also be considered. Saucer-shaped deformation Another effect that can become problematic with increasing power density is the saucer-shaped deformation of the pressure plate, also known as thermal shielding [21, 22]: As a result of the unilaterally converted frictional power during a braking process, the pressure plate heats up and = 2 = Science and Research 24 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 The spring force is applied by a variable number of compression springs and is approximately constant due to the small pitch of spring required. The average friction radius would be mathematically the average radius at the lining ring averaged over the surface. In reality, however, it has been shown that the effective friction radius wanders and so that the braking torque is also uneven [16]. The friction coefficient depends mainly on the friction partners, but it is also influenced by other parameters such as temperature, contact pressure or sliding speed. Commercially available spring-applied brakes usually have two friction surfaces. Organic friction linings In the case of organic friction linings, the carrier matrix of the friction lining consists of elastomers or synthetic resins into which fillers with various functions are introduced. These are abrasives to increase the coefficient of friction, lubricants to reduce wear and fibres to increase strength and thermal resistance. The temperature resistance of the carrier matrix and the fibres significantly limits the thermal load capacity of the friction lining. The coefficient of friction collapses under thermal overload. Organic friction linings are used in clutches and springapplied brakes, among others, in spring-applied brakes often with a high rubber content. [16,17,18] The coefficients of friction are usually given for sliding speeds up to a maximum of 20 m/ s; data for higher sliding speeds are rarely found. Friction systems with organic friction linings form a friction layer on the surface during wearing in, which significantly influences the behaviour of the friction coefficient. The generally low thermal conductivity of organic friction materials leads to a high temperature at the friction surface, especially when they run dry, i.e. are enclosed by the insulator air. Due to these short time high temperatures of locally up to 800°, the highly carbonaceous friction layer with a thickness of 0.001 - 20 µm is formed. This friction layer has a decisive influence on the behaviour of the friction value; at the same time, it serves as a thermal protective shield for the friction lining underneath. If the friction lining runs in under constant thermal load, the friction layer stabilises after a certain number of braking operations and thus also stabilises the coefficient of friction. If the thermal load changes, the behaviour of the friction layer and thus the coefficient of friction also changes. [17, 18, 19] The model according to SEVERIN, D. [18] considers the surface elements of the friction layer: Here, a surface element participates in the friction force transmission until it is damaged. Then it regenerates and after a certain period of time it starts to wear again. If damage and regeneration of surface elements are in equilibrium, the coefficient of friction remains constant. If the thermal load increases unacceptably, more surface elements are damaged than regenerated. The result is that the coeffiexpands on one side. If thermally induced stresses in the pressure plate exceed the yield point of the material, irreversible plastic deformation occurs. The decrease of the yield point at high temperatures additionally favours the plastic deformation. Figure 2 shows the resulting shape of a plate that is inclined towards the friction surface. Such a set-up entails various problems: Compared to the non-saucer-shaped state, the air gap is reduced when the brake is open. On the one hand, this can lead to unwanted dragging of the brake; on the other hand, the compensation of the reduced gap requires an increase of the actuation travel, which is by design not always possible. In the braked state, the saucer-shaped deformation leads to changed friction conditions: As a result of the change in geometry, instead of a friction surface, there is cantilevering in the outer area of the friction surface provided by the design. This causes an increase in the friction power density and can lead to thermally induced friction lining damage, breakage of friction lining due to increased surface pressure, uneven wear and noise development or metallurgical changes [10]. Due to the shift in the effective friction radius, the braking torque deviates from the brake’s design braking torque. The thermal build-up of individual components of brakes is well known and has already been mentioned several times in the literature: B ESTLE , H. et al. [21] describes thermal distortion on armature discs and flange surfaces of spring-applied brakes and attributes this to unilateral heating, which causes unilateral expansion of the components. As a proposed solution, B ESTLE , H. et al [21] lists a radial reduction of the friction surface, which, however, appears to be practicable only to a limited extent due to the high friction power density. B REUER , B. et al [23] describe the thermal build-up of unilaterally bolted brake discs of vehicle brakes, which can be avoided by allowing radial expansion using an axially floating bearing. A UDEBERT , N. et al [24] have investigated the saucershaped deformation of clutch discs in automatic transmissions and deduced that it depends on a dimensionless geometric shape factor that can be used to predict the tendency of a saucer-shaped deformation as a function of temperature differences at the brake disc. X IONG , C. et al [22] have developed a calculation method for determining the critical temperature-related moments in the pressure plate based on the curved beam model. So far, however, the literature does not show any solution approaches with which the tendency to saucer-shaped deformation can be reduced for a given geometry and braking energy. Objectives and procedure In the state of the art section, the problems of organic friction linings and pressure plates in high-speed brakes Science and Research 25 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 Figure 2: Problems due to saucer-shaped deformationa brake open, b brake closed inertia to be braked can be represented by the drive, but in this configuration the maximum braking torque is limited to the motor torque multiplied by the installed gear ratio. Therefore, a flywheel mass module can be installed as an option, which can represent load inertias at large braking torques. This is followed by the brake module with the connection for the brake to be examined and the measurement technology. Alternatively, the brake can also be attached directly to the gearbox output of the drive module. The test stand data are listed in Table 1. Investigation of the friction torque development of organic friction linings at sliding speeds up to 50 m/ s The load on the friction components during braking is influenced by various parameters, such as speed/ sliding speed, braking torque, surface pressure, friction work and ambient temperature. To get an overview of the possible range of use in high-speed applications, the parameters speed from which braking takes place and friction work are varied. This also adjusts the friction power. In conjunction with the sliding speed and the specific friction work, comparability with other geometries can be established. The remaining influencing parameters are kept constant by always using the same brake in the same configuration and the laboratory environment. The brake data are listed in Table 2, the speeds to be investigated are specified in Table 3. Science and Research 26 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 during operation were highlighted. The aim of this paper is to investigate to what extent higher sliding speeds can also be achieved with organic friction linings and what measures can be taken to reduce the saucer-shaped deformation of pressure plates. This is done as follows: - Description of the test technique - Investigation of the friction torque development of organic friction linings on the basis of a spring-applied brake at sliding speeds of up to 50 m/ s under various loading situations. - Derivation of the physical background that leads to the build-up of pressure plates - Development of approaches to reduce the uplift - Experimental investigation of the approaches found - Conclusion Test engineering Figure 3 shows the structure of the modular brake highspeed friction test stand in the form of a schematic diagram. A synchronous servo drive is selected as the drive to provide constant and sufficient dynamics and torque over a wide speed range. In order to achieve the desired sliding speed of up to 100 m/ s in the friction contact, a gearbox is connected downstream of the servo drive; there is a choice of two gear ratios, i = 0.45 and i = 0.22. By default, any load Figure 3: Schematic diagram of brake high-speed friction test rig Drive data: Rated torque 95 Nm Rated speed 4,500 rpm Gearbox 1 Gear ratio 0.45 Output speed max. 10,080 rpm Output torque 42 Nm Gearbox 2 Gear ratio 0.22 Output speed max. 20,000 rpm Output torque 21 Nm Measurement data: Torque, speed, braking force, temperature, axial displacement measurement, current, voltage Table 1: Test bench data The definition of the friction work is based on the definition of VDI guideline 2241-1 [25]. This defines the permissible friction work Q zul for switchable, externally actuated friction clutches and brakes at which thermal overload does not occur. Q zul depends on the switching frequency and the permissible friction work for a single switching Q E . This value depends on the size of the brake; for the spring-applied brake used, Q E is 24,000 J. For each speed, braking is carried out at 5, 12.5 and 50 percent of Q E . This is set via the simulated inertia. Before the start of the tests, the brake was broken in in order to establish a two-dimensional contact between the friction partners and to enable the formation of the friction layer [19]. Figure 4 shows the torque and speed curves of a representative braking from 1,500 rpm and 10,000 rpm each at 12.5 % (lower graph) and 50 % Q E (upper graph). It can be seen that for the braking at 12.5 % Q E , the braking torques at 1,500 rpm (blue) and 10,000 rpm (red) are very similar at about 37.5 Nm. The high Science and Research 27 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 Nominal braking torque 32 Nm mean friction radius 46.25 mm max. mean sliding speed 48.4 m/ s Friction surface of one lining ring 6,527 mm² Q E 24,000 J Table 2: Brake data Braking speed in rpm Average sliding speed in m/ s Maximum specific nominal friction power in W/ mm² Specific friction work in J/ mm² for 5% Q E / 12,5% Q E / 50% Q E 100 0.5 0.03 0.09 / 0.23 / 0.92 Ordinary braking torque mean speed 1,500 7.3 0.39 Usual operating speed 3,000 14.5 0.77 Usual operating speed 6,700 32.5 1.73 High-speed braking 10,000 48.4 2.58 High-speed braking Figure 4: Torque and speed curve for braking from 1,500 rpm (blue) and 10,000 rpm (red) with 12.5 % (bottom) and 50 % QE (top) respectively Table 3: Examined speeds/ sliding velocities for organic friction linings an initial sliding speed of 32 m/ s. However, it becomes clear that there are areas where the commercial brake works well even at high sliding speeds. In addition to the torque drop, friction lining adhesions were visible on the counter friction surfaces during braking from high speeds with high energies. This indicates that the friction lining or at least its components were briefly thermally overloaded. In addition, a slight saucershaped deformation of the armature disk and brake flange were observed. No other mechanical damages to the rotor or friction lining due to e.g. centrifugal force were observed. Physical background of the saucer-shaped deformation During the braking process, the pressure plate is heated on one side of the circular friction surface. At high friction power, a lot of energy is applied in a very short time, resulting in a high temperature gradient between the friction surface and the surrounding material (Figure 6 a and b). This causes the material in the area of the friction surface to expand much more than the surrounding material. The pressure plate gets a saucer shape with the convex surface on the friction side. The obstruction of the expansion causes compressive stresses in the heated area and tensile stresses in the surrounding area. If the compressive stresses exceed the yield point of the material, which is reduced due to the high temperature, plastic Science and Research 28 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 friction power is apparently well tolerable for small inertias, i.e. for short friction times and the associated low friction work. When braking with 50 % Q E , the braking torque of 10,000 rpm is lower from the start (33 Nm) and drops even further to about 28 Nm at the beginning of braking. Only when the speed drops significantly towards the end of braking does the braking torque rise to about 35 Nm, which is close to the average value for all braking operations. Due to the high inertia, the friction power is at a high level for a longer time period, which seems to have caused the temperature to become too high, at least in the first half of the braking process. This probably damaged the friction layer, which leads to a drop in the friction coefficient and thus the braking torque. However, according to Severin [18], the torque increase at the end of braking suggests a “regeneration of the friction layer”. Figure 5 shows the determined characteristic diagram of the tested brake, in which the friction torque is plotted against sliding speed and friction work. The diagram represents the mean value of five braking operations per operating point. It depicts that at 5 % and 12.5 % QE the braking torques are at the same level across speeds. For braking at 50 % Q E , the braking torque of the high-speed braking drops significantly, especially in comparison to braking at usual speeds and also in comparison to braking with less inertia. The braking torque drops even more significantly at an initial sliding speed of 48 m/ s than at Figure 5: Braking torques shown as a function of sliding speed and friction work compression occurs in the area of the friction surface. In the surrounding area, plastic strain occurs to a lesser extent due to the lower temperature. After cooling, a saucer-shaped deformation remains with the concave surface on the friction side, because the plastically compressed material contracts more than the uncompressed or stretched material in the surrounding area (Figure 6 c). Tensile stresses remain in the compressed area and compressive stresses in the surrounding area. Theoretically, the saucer-shaped deformation should not increase significantly after the first braking, because the thermal expansion from the second braking on initially only reduces the pre-tension in the pressure plate and would secondly compensate for the compression. Only beyond this point does an excess of expansion occur again. However, in a real friction system, a constant homogeneous energy input to the friction surface cannot be assumed, thus changes can still be expected after the first braking. Development of approaches to reduce the saucer-shaped deformation The high stresses resulting from thermal expansion are the main cause for the plastic saucer-shaped deformation. In principle, there are three measures that can, given an unchanged friction lining geometry, reduce the plastic saucer-shaped deformation: - Reducing the thermal expansion - Reducing the stresses caused by expansion - Increasing the stresses tolerable without plastic deformation. In the following, based a reference pressure plate, geometric and material-technical approaches are discussed by which these measures can be implemented. Reference pressure plate The full-surface geometry shown in Figure 7 is used as the reference pressure plate. There are guide noses on the outer diameter to accommodate it in the brake housing. Table 4 summarises the properties of the reference pressure plate. The standard steel S355 with medium strength was selected in order to obtain a cost-effective reference pressure plate. The frequently used cast iron lamellar graphite has an even lower saucer-shaped deformation resistance (see table 5) and was therefore not used as a reference. This means that S355 achieves a good compromise between low saucer-shaped deformation resistance and low costs. Science and Research 29 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 Figure 6: Phases of the saucer-shaped deformation process - a during heating, b after heating, c after cooling In the following considerations, the reference pressure plate is compared to the modified pressure plate shown in Figure 8. This has 16 radial slots, each of which is inclined 15° in the direction of rotation of the rotor in order to reduce shearing of the friction lining at the edges. Investigation of material engineering approaches The choice of a suitable material can also assist in implementing the measures for lowering plastic deformation. The identification of suitable materials is aided by an analogy with thermal fatigue strength, which has been documented in the literature, for example, by B ÜRGEL , R. et al [27] in the context of high-temperature materials. There, the considerations of the three measures mentioned are combined in a common consideration. B ÜRGEL , R. et al [27] understand the temporary and locally limited introduction of a high thermal energy into a component, which leads to a high temperature gradient within the component, as a so-called thermal shock. The different local temperatures within the component result in high stresses in the component, which can lead to thermal fatigue damage if the thermal stress is repeated cyclically. To evaluate materials with respect to their thermal shock sensitivity, [27] introduces the thermal stress index χ (4). Science and Research 30 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 Investigation of geometric approaches The first measure, reducing thermal expansion, can, from a geometric point of view, be achieved by stiffening the structure. Such stiffening can be accomplished by increasing the pressure plate thickness as well as reducing the inner diameter or increasing the outer diameter. It should be noted that such geometry adjustments lead to additional strain restraint and to larger temperature gradients during heating and may therefore prove counterproductive. Lowering of strain-induced stresses can be obtained by reducing the pressure plate thickness and the radial installation space, as proposed for example by Z AGRODSKI [26] for plates of a clutch. However, both measures contradict the objective of reducing thermal expansion and can therefore also have negative consequences for thermal expansion. Due to the opposing influence of different effects, such adjustments of the pressure plate are only recommended after extensive investigations. Another way to reduce strain-induced stresses is to allow an unhindered thermal expansion. In the case of the present disc geometry, thermal expansion leads in particular to high tangential stresses. The tangential stresses can be reduced by radial slotting from the inside diameter of the friction surface to the outside diameter of the pressure plate. Dimension Value Outer diameter 140 mm Inner diameter 72 mm Thickness 5 mm Outer friction diameter 139 mm Inner friction diameter 90 mm Mean friction radius 57,25 mm Material Structural steel S355 Table 4: Properties of the reference pressure plate Figure 7: Reference pressure plate made of S355 Figure 8: 16-slotted pressure plate made of S355 (4) Legend: χ Thermal stress index R m Tensile Strength λ Thermal conductivity α Coefficient of thermal expansion E E-modulus A material is particularly sensitive to thermal shock if there is a high temperature gradient in the material due to a low thermal conductivity λ. A high coefficient of thermal expansion αcauses large strains which, with a large E-modulus, result in high stresses within the component. This behavior must be considered in relation to the tensile strength of the material R m to determine if the thermally induced stress is critical in terms of potential damage. In the case of thermal saucer-shaped deformation of static compression plates, the thermal conductivity λ, the coefficient of thermal expansion α, and the E-modulus lead analogously to high stresses in the component and thus to severe saucer-shaped deformation. However, since the cause of thermal saucer-shaped deformation is plastic deformation, the yield strength RP must be used as material reference. From this, the saucer-shaped deformation resistance A (5) can be derived according [28]: (5) Legend: A saucer-shaped deformation resistance R p Yield strength λ Thermal conductivity α Coefficient of thermal expansion E E-modulus = R m E A = R p E Table 5 gives an overview of the saucer-shaped deformation resistance A for exemplary materials. According to this, E360AR structural steel is particularly sensitive to thermal saucer-shaped deformation, while little thermal saucer-shaped deformation is to be expected with copper. In reality, however, the choice of a suitable material usually involves compromises regarding function and cost. In the experimental investigations documented below, four pressure plates are compared: • Reference pressure plate made of S355 steel • Slotted pressure plate made of S355 steel • Slotted pressure plate made of 42CrMo4 steel • Slotted pressure plate made of CW004 copper Experiment In order to verify the above approaches, five successive brake applications per pressure plate were performed according to the specification shown in Table 6. After a short low-energy braking-in, five successive braking operations were conducted, each starting at a speed of 14.000 rpm. The plates were subjected to the energy and frictional power specified in Table 3. The cooling phase between the brakings was selected so that the temperature of the pressure plate fell below 70 °C. The saucer-shaped deformation of the pressure plates was measured after the second, fourth and fifth braking operation. The distance to the inner radius of the friction surface and the outer radius of the friction surface were measured at three points of the friction surface at the flat-lying pressure plate using a laser triangulation sensor. The difference between the two distance values is the saucer-shaped deformation at this point. For the over- Science and Research 31 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 g y g Parameter Value Friction energy per friction surface 55.5 kJ Maximum friction power per friction surface 43 kW Braking torque per friction surface 30 Nm Speed 14,000 rpm Mass moment of inertia to be braked per friction surface 51,750 kgmm² Table 6: Data of a braking cycle to investigate the saucer-shaped deformation Material group Material saucer-shaped deformation resistance A Source Cast iron lamellar graphite EN-GJL-200 6.8 [29a, 30] Cast iron nodular graphite EN-GJS-700-2 6.2 [29b] Cast steel G10MnMoV6-3 10.3 [31a] Structural steel S355JR 8.4 [32] E360AR 3.4 [31b] Case-hardening steel 16MnCr5 9.1 [31c] Heat-treatable steel 42CrMo4 10.0 [31d, 33] Copper (99.9%) CW004A 21.4 [34] Table 5: saucer-shaped deformation resistance of exemplary materials Figure 10 shows the slotted pressure plate made of CW004A lying next to the S355 reference pressure plate on a measuring plate. On the copper pressure plate, there is hardly any visible saucer-shaped deformation, while in the case of the reference pressure plate made of S355 it is clearly visible to the naked eye. Conclusion and outlook In principle, it can be said that the commercially available friction system investigated can deliver the nominal braking torque even at high sliding speeds, as long as the friction work does not become too great. Thus, the spring-applied brake offers the potential to be used in the application scenarios outlined at the beginning without necessarily having to resort to higher-quality, more expensive friction materials. However, if the permissible scope of application is to be expanded, it is necessary to determine the definition of the permissible friction work as a function of sliding speed, because this takes the frictional power that causes temperature peaks at the onset of braking into account. Science and Research 32 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 all assessment, the mean of the three measured values was calculated. The measured values were validated with a feeler gauge measurement. The test results are shown in Figure 9. Qualitatively, they confirm the preliminary considerations. Also, as expected, a large part of the saucer-shaped deformation occurs after only a few braking operations. At 0.42 mm, the reference pressure plate made of S355 exhibits the highest saucer-shaped deformation after five braking operations. The second-highest saucer-shaped deformation is exhibited by the slotted S355 pressure plate with 0.23 mm. Consequently, the saucer-shaped deformation was almost halved by the slotted geometry. The change in material to the higher-strength steel 42CrMo4 again cut the saucer-shaped deformation in half, with a buildup of 0.1 mm. As expected, the copper pressure plate showed the least deformation with 0.04 mm. The small amount of saucer-shaped deformation indicates that the yield stress of the material is not reached over a large area in this pressure plate. The slight deformation is presumably due to local hotspots, through which the yield stress is exceeded selectively. Figure 10: Slotted pressure plate made of CW004 (left) and reference pressure plate made of S355 (right) lying on a measuring plate Figure 9: Test results of the saucer-shaped deformation The experimental investigation has confirmed that the solutions presented to reduce thermal deformation are effective. By introducing radial slots, it was possible to halve the buildup for identical material. Further reductions in the saucer-shaped deformation could be achieved by selectively adjusting the material on the basis of the saucer-shaped deformation index. It was also shown that the deformations can be further reduced with an increasing saucer-shaped deformation index, meaning there is a qualitative relationship between the saucer-shaped deformation index and deformation. In addition to the approaches presented here for reducing plastic deformation, more advanced approaches are also conceivable: For example, by training the pressure plates prior to the initial braking process, a stress state could be specifically set that compensates for the resulting stresses causative for the deformation. Pressure plates whose initial geometry is opposite to the resulting saucer-shaped deformation (tapered friction surface) could be deformed by the saucer-shaped deformation in such a way that their friction surfaces are approximately flat. Both approaches require precise knowledge of the expected load and saucer-shaped deformation and can therefore require adequate elaborate preliminary investigations prior to implementation. In summary, it can be stated that the saucer-shaped deformation of pressure plates that are subjected to a high energy input on one side can be significantly reduced by introducing radial slots. The effect of the slots is that thermal expansion due to the high energy input is possible, thus reducing stresses. This should be the first measure to be applied in practice. Further reductions in saucer-shaped deformation can be achieved by an adapted material selection. Since steel materials are predominantly used in practice, the recommendation is to select a material with higher strength while maintaining good thermal conductivity. Literature [1] Schlecht, B.: Maschinenelemente 1 - Festigkeit, Wellen, Verbindungen, Federn, Kupplungen, München: Pearson Studium, 2005 [2] Eitel, L.: (2016) „Trends in clutzches and brakes leverage software and customization. MOTION CONTROL TIPS“, https: / / www.motioncontroltips.com/ trendsclutchesbrakes-leverage-software-customization, 2016, Zugriff am 02. März 2022 [3] Langer, G.: „Leicht und dennoch leistungsstark - Elektrisch schaltbare Haltebremse für Serienanwendungen“, IEN D-A-CH: 2017. https: / / www.ien-dach.de/ artikel/ leicht-und-dennochleistungsstark. Zugriff am 02. März 2022 [4] Yokohama, T. A.: „Friction Brake and Vhicle-Mounted Apparatus“, Europäische Patentschrift EP3614010A1, 2020 [5] Binder, A.; Schneider, T.: „High-speed inverter-fed AC drives“, International Aegean Conference on Electrical Machines and Power Electronics, Bodrum: 2007 [6] Mekuria, Y. G.: „Development of a High Speed Solid Rotor Asynchronous Drive fed by a Frequency Converter System“, Dissertation TU Darmstadt, 2013 [7] Centner, M.: „Entwurf und Erprobung schnelldrehender Asynchronmaschinen unter besonderer Berücksichtigung der magnetisch aktiven Materialien“, Dissertation TU Berlin, 2009 [8] Pasquarella, G. 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Tribology Transactions“, Tribology Transactions, Volume 65, Issue 5, 2022 [15] Kiel, E.: „Antriebslösungen“, Berlin, Heidelberg: Springer-Verlag, 2007 [16] Chr. Mayr GmbH + Co. KG, „Elektromagnetische Sicherheitsbremsen“, Süddeutscher Verlag, 2014 [17] Bergheim, M.: „Organisch gebundene Kupplungsbeläge - Möglichkeiten und Grenzen“, Kupplungen in Antriebssystemen 1997, Verein Deutscher Ingenieure: VDI-Tagung, Fulda, 3. und 4. März 1997, VDI Verlag [18] Severin, D.: „Zum Verständnis der Reibsysteme in trockenlaufenden Bremsen und Kupplungen“, Kupplungen in Antriebssystemen 1997, Verein Deutscher Ingenieure: VDI-Tagung, Fulda, 3. und 4. März 1997, VDI Verlag [19] Musiol, F.: „Erklärung der Vorgänge in der Kontaktzone von trockenlaufenden Reibpaarungen über gesetzmäßig auftretende Phänomene im Reibprozess“, Dissertation TU Berlin, 1994 [20] Albers, A.; Ott, S.; Schepanski, N.: „Ermittlung der Leistungsgrenze trockenlaufender Friktionspaarungen in Abhängigkeit der Belastungsparameter“, Kupplungen und Kupplungssysteme in Antrieben 2015: VDI-Fachtagung mit Fachausstellung, Karlsruhe, 28. und 29. April 2015, VDI Verlag Science and Research 33 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 [29b] Meuselwitz Guss Eisengießerei Werkstoffkenndaten - Späroguss: Gusseisen mit Kugelgraphit. https: / / www.meuselwitz-guss.de/ fileadmin/ daten/ Dateien/ pdf/ werkstoffe/ Werkstoffkenndaten_Kugelgra phit_getrennt_gegossen.pdf. Zugriff am 03. März 2022 [30] Baharin, B.; Akop, M., Z.; Salim, M., A.; Mansor, M., R.; Saad, A., M.; Rosli, M., A., M.; Arifin, Y., M.; Tahir, M., M.: „Thermal stress effect on disc brake rotor for NGV vehicle“, Melaka: Proceedings of Mechnical Engineering Research Day, 2017 [31a] Make it from database EN 1.5410 (G10MnMoV6-3) Cast Steel. https: / / www.makeitfrom.com/ material-properties/ EN-1.5410-G10MnMoV6-3-Cast-Steel. Zugriff am 03. März 2022 [31b] Make it from database EN 1.0070 (E360) Non Alloy Steel. https: / / www.makeitfrom.com/ material-properties/ EN-1.0070-E360-Non-Alloy-Steel. Zugriff am 03. März 2022 [31c] Make it from database EN 1.7131 (16MnCr5) Chromium Steel. https: / / www.makeitfrom.com/ material-properties/ EN-1.7131-16MnCr5-Chromium-Steel. Zugriff am 03. März 2022 [31d] Make it from database EN 1.7225 (42CrMo4) Chromium Steel. https: / / www.makeitfrom.com/ material-properties/ EN-1.7225-42CrMo4-Chromium-Molybdenum-Steel. Zugriff am 03. März 2022 [32] Da Silva, L., S.; Lamas, A.; Jaspart, J., P.; Bjorhovde, R.; Kuhlmann, U.: „Fire Design of steel structures“, Berlin: Wilhelm Ernst & Sohn Verlag für Architektur und technische Wissenschaften, 2012 [33] Emde, T.: „Mechanisches Verhalten metallischer Werkstoffe über weite Bereiche der Dehnung, der Dehnrate und der Temperatur“, Dissertation, Rheinisch-Westfälische Technische Hochschule Aachen, 2008 [34] Deutsches Kupferinstitut Cu-ETP: https: / / www.kupfer institut.de/ wp-content/ uploads/ 2019/ 11/ Cu-ETP.pdf. Zugriff am 07. März 2022 Science and Research 34 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0004 [21] Bestle, H.; Fleichmann O.; Hecht, M.; Huber, J.; Kempf, P.; Unsin, K.: „Elektromagnetische Sicherheitsbremsen- Auslegung, Ansteuerung, Bauformen“, München: Verlag Moderne Industrie, 2014 [22] Xiong, C.; Ma, B.; Li, H.; Zhang, F.; Wu, D.: „Experimental Study and Thermal Analysis on the Buckling of Friction Components in Multi-Disc Clutch“, Journal of Thermal Stresses, 38-11, 2015 [23] Breuer, B.; Bill, K., H.: „Bremsenhandbuch - Grundlagen, Komponenten, Systeme, Fahrdynamik“, Berlin: Springer, 2012 [24] Audebert, N.; Barber, J., R.; Zagrodzki, P.: „Buckling of automatic transmission clutch plates due to thermoelastic/ plastic residual stresses“, Journal of Thermal Stresses, vol. 21, iss. 3-4, 1998 [25] Verein Deutscher Ingenieure: VDI-Richtlinie 2241-1: „Schaltbare fremdbetätigte Reibkupplungen und -bremsen“, VDI-Verlag, 1982 [26] Zagrodzki, P.; Truncone, S.: „Generation of hot spots in a wet multidisk clutch during short-term engagement“, Wear, vol. 254, iss. 5-6, 2003 [27] Bürgel, R.; Maier, H., J.; Niendorf, T.: „Handbuch Hochtemperatur-Werkstofftechnik: Grundlagen, Werkstoffbeanspruchungen, Hochtemperaturlegierungen und -beschichtungen“, Wiesbaden: Vieweg+Teubner, 2011 [28] Schadomsky, M.; Blumenthal, L.; Zimmer, D.; Peter, S., Boros, L.: „Maßnahmen zur Reduzierung von plastischer Verformung an statischen Druckplatten von Lammellenbremsen infolge einseitig eingebrachter hoher Reibleistung“, Forschung im Ingenieurwesen 86, 891- 901, 2022 [29a] Meuselwitz Guss Eisengießerei Werkstoffkenndaten - Grauguss: Gusseisen mit Lamellengraphit. https: / / www.meuselwitz-guss.de/ fileadmin/ daten/ Dateien/ pdf/ werkstoffe/ Werkstoffkenndaten_Lammel lengraphit.pdf. Zugriff am 03. März 2022 Introduction Crossed helical gear units with plastic gears are the current state of the art in positioning and actuating drives. The range of plastic materials with their individual material properties offers the potential to meet different application requirements safely and reliably. In addition to the geometric parameters and for the design of gears, the properties of the flank contact must be described in detail to meet specific requirements. Previous approaches to calculate contact behavior on a tribological level have concentrated on two contacting cylindrical gears. The operating behavior of a steel-plas- Science and Research 35 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 Local considerations and experimental results on the contact behavior of crossed helical gears with general flank geometries Linda Becker, Peter Tenberge* submitted: 10.10.2023 accepted: 21.03.2024 (peer-review) Presented at the GfT Conference 2023 Schraubradgetriebe gehören zu Zahnradgetrieben mit gekreuzten Radachsen und bestehen in den meisten Anwendungen aus einer Stahl-Schnecke und einem Kunststoff-Rad. Aufgrund der hohen Übersetzung auf kleinem Bauraum werden sie in Stell-, Nebensowie in Positionierantrieben eingesetzt. Fortschritte in der Forschung zu Hochleistungspolymeren ermöglichen zunehmend leistungsstärkere Verzahnungen, die in Abhängigkeit der Eigenschaften des Kunststoffmaterials, an die Belastungsgrenzen von Stahlwerkstoffen gelangen können. Forschungen zu Schraubradgetrieben der Paarung Stahl/ Kunststoff mit geometrisch optimierten Zahnflankengeometrien bieten Effizienzsteigerungen und nachweislich Wirkungsgradpotentiale, wodurch höhere Tragfähigkeiten sowie ein Anstieg in der Lebensdauer realisierbar werden. Eine Aufschlüsselung des Kontaktverhaltens im Eingriffsgebiet ermöglicht Betrachtungen von Reibenergien und stellt die Basis für örtliche Verschleiß- und Tragfähigkeitsbewertungen. Auf dieser Grundlage lassen sich effizientere Schraubradverzahnungen mit allgemeinen Flankenformen gestalten, die infolge der Lebensdauersteigerung einen positiven Einfluss auf die Ressourcenschonung und Nachhaltigkeit haben. Schlüsselwörter Schraubradgetriebe, Freiformgeometrien, ZC, Reibung, Temperatur, Kontaktverhalten Crossed helical gears belong to gear units with crossed wheel axles and in most applications consist of a steel worm and a plastic wheel. Due to the high transmission ratio in a small installation space, they are used in actuators, auxiliary drives and positioning drives. The research into high-performance polymers is accompanied by increasingly powerful gears, depending on the properties of the plastic material, to reach the load limits of steel materials. Research on steel/ plastic crossed helical gears with geometrically optimized tooth flank geometries offers efficiency improvements and demonstrable efficiency potential, enabling higher overall load capacities and an increase in service life. A more detailed resolution of the contact behavior in the contact area allows friction energies to be considered and provides the basis for local wear and load capacity evaluations. Based on this, new crossed helical gears with general flank geometries can be designed, which have a positive influence on resource conservation and sustainability as a result of the increase in lifetime. Keywords Crossed helical gears, ZC-geometry, Friction, Temperature, Contact behavior Kurzfassung Abstract *Dr.-Ing. Linda Becker Orcid-ID: https: / / orcid.org/ 0000-0002-2183-128X Prof. Dr.-Ing. Peter Tenberge Chair of Industrial and Automotive Drivetrains Ruhr-University Bochum, D-44801 Bochum, Germany contact ellipse. The general calculation case according to [ISO17], which is recommended for gears with an axis crossing angle, such as bevel gears, converts these to virtual spur gears with modification parameters, so line contacts of the same material combination with twodimensional contact behavior are still present as the basis of the calculation. Regarding this, the assumptions according to [ISO17] are not transferable to crossed helical gears, so the development of the equations are a current research task. The consideration of the entire contact area represents an extension of the previous calculation approaches. Theoretical consideration of contact behavior For the calculation approach of ISO 6336-20 [ISO17], first it is necessary to describe the orientation of the contact ellipse on the tooth flanks and the individual velocity components. The calculation according [BECK23b] considers the contact behavior along the path of contact as a function of a parameter u, which runs from the worm root to the tip and for the wheel in the opposite direction. The contacting tooth flanks move in a certain angular position on a path through the contact ellipse, while a part of this path being sliding through. In each contact position, the velocities and directions of movement are different for general flank geometries on worm and wheel. The following section is intended to derive these parameters and to find the equations according to [ISO17] for the application of crossed helical gears for a steelplastic material combination with point contact. General curved surface structures also result in new contact geometries, so the assumption of two equivalent contacting cylinders are no longer permissible. The curvatures are reduced to two substitute ellipsoids in contact [BECK23b]. For general flank shapes, the contact ellipse is arranged in different angular positions in relation to the radial section of the gear. Deviating from the involute it is no longer permissible to assign the major semi axis a H to the tooth width direction. The position of the contact ellipsoid semi-axes has to be considered individually in a case distinction for each contact position and has an influence on the calculation [BECK23b]. The contact ellipse moves along a general curved contact path, as illustrated in the following ZC example (Figure 1). According to Boehme [BOEH20], the velocities are considered in individual coordinate directions. Sliding velocities in gears with crossed axes are composed of a sliding component in the tooth height direction and a component in the tooth width direction. The second one is also called helical sliding and increases with increasing axis crossing angle. The surface stress, which leads to frictional wear on the plastic gears, takes place under Science and Research 36 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 tic pairing and a three-dimensional contact requires new calculation approaches. Recent research results [BECK23b] provide a computational algorithm that represents the contact behavior of crossed helical gears in the contact area at each contact position. In addition, this algorithm allows the variation of flank shapes, which are described by a mathematical polynomial. As a result, non-involute geometries with potentials in terms of load carrying capacity and gear life are obtained. Based on the new approach for general flank geometries, it is possible to analyze the contact behavior with regard to tribological influences in flank contact in more detail. Particularly in the case of crossed helical gears with a design-related high sliding contact on the gear surfaces, there is a research task regarding the frictional-energetic interactions in the design process. State of research The current state of research on crossed helical gears provides detailed considerations of the geometric relations in the design process as well as a description of flank pressures, efficiencies and sliding paths in the entire contact area [BECK23b]. The sliding-rolling behavior with its tribological parameter is important for the load carrying capacity, especially for plastic gears, and has not been considered very much so far. Pech [PECH11] has carried out wear investigations on practical gears for a reference gear set and described correlations for ZI geometries which reflect the maximum wear at the screw point for his geometries and material pairing. The procedure is based on geometric correlations and operating properties of worm gears, which show different effects in their operating behavior. It is shown that design criteria of crossed helical gears do not necessarily have to correspond to those of worm gears [BECK23b]. Sucker [SUCK13] has taken Pech’s approach and extended it with a thermal network. Both calculations are less focused on tribological causes for the corresponding wear amounts. In crossed helical gears, essential influences on the contact behavior are composed of the frictional effects of the contact partners in contact. This is accompanied by the specific energy transfer capacity for dissipating the contact temperature from the gear mesh. To describe these parameters, ISO 6336-20 [ISO17] deals with the calculation of the scuffing load capacity of cylindrical steel gears. Scuffing damages in steel-plastic material combinations are rare, but temperature and friction considerations have a decisive importance. The ISO guidelines refer specifically to cylindrical gears of the steel-steel material combination with a semi-elliptical flank contact, in which the contact zone is simplified to a band-shaped surface with twice the width of the minor semi-axis b H of the the sliding velocity along the path of contact in the direction of the sliding motion. To determine the sliding path orientation on the tooth flanks, the directions of the tangential velocities v t1 and v t2 of the worm and wheel are derived with reference to a coordinate system at the contact point. The geometric relationship is shown in Figure 2. The tangential velocities v t1,2 at the contact point E show the difference between the absolute and normal velocities. The indices “1” and “2” indicate the affiliation to the worm and the wheel. The velocities can be divided in the yand z-direction of the coordinate system located at the point of contact to determine the angular position. The direction results from corresponding tangential vectors [BECK23b]. As a reference position, between the angles β vt1 and β vt2 there is a horizontal auxiliary line parallel to the x-y plane through the contact point. The angles β vt1 and β vt2 are calculated according to equation (1) and (2). (1) (2) The angle β Ha indicates the orientation of the major semi-axis a H to the horizontal auxiliary line. According to this, it is possible to transfer the consideration of the tangential velocity directions for worm and wheel to the position of the major semi-axis, according to the procedure in [ISO17]. = tan ( ) ( ) = ( ) ( ) Science and Research 37 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 Path of contact Contact ellipse u Path of contact Contact ellipse u Contact ellipse u Path of contact Start of contact End of contact Middle contact are Figure 1: Change of orientation of the contact ellipse along a ZC path of contact n q(u) v t1 v t2 β vt2 β vt1 Horizontal Line parallel to the x-y plane E Tangent vector 1 at the generating profile Tangent vector 2 at the generating profile termv2z(u) termv1z(u) termv1y(u) termv2y(u) xt2q u ( ) yt2q u ( ) zt2q u ( ) xt1q u ( ) yt1q u ( ) zt1q u ( ) v t1 v t2 E a H b H Contact ellipse β Ha β vt1 β vt2 n q(u) a) b) Figure 2: Derivation of the angular positions of the tangential vectors on the worm for the tangential velocities v t at the contact point E, a) representation of the vector components, b) related to a contact point E Path of contact β Ha Horizontal auxillary line X Y Z a H s ellipse2 b H φ slide2 Tangential velocity worm v t1 β vt2 The sliding path of the contact partners through the contact ellipse is critical in terms of friction energy. Due to the three-dimensional velocity conditions, the path s ellipse is neither equally oriented nor equally long for worm and wheel, so the consideration must be carried out separately. The directions of the sliding motions φ slide1,2 result from the consideration of the orientation β vt1,2 of the tangential velocities as well as the angle β Ha to the major semi-axis a H . Figure 3 illustrates the relationships between the parameters. Figure 3: Orientation of tangential velocities and sliding movements in the radial section of the wheel B M and the contact time of the heat source T k . Equation (5) describes the increasing flash temperature θ fl according to the approach from Blok for local point contacts of crossed helical gears. (5) The maximum contact temperature at the tooth flank surface θ max results from the constant bulk temperature θ m and the flash temperature θ fl , which depends on the location on the contact path. (6) The specific contact energy E k describes the temperature-independent frictional energy input that finally leads to the heat in the contact point [LOOS15]; [STUH22]. The distribution of the partial heat flows is carried out under the assumption that equal contact temperatures result at the contacting surfaces. For crossed helical gears, the distribution of the partial heat fluxes is unevenly distributed among the contacting partners, since the steel worm has a better thermal conductivity than the plastic gear. The coefficient γ Ek describes the distribution of the heat flux among the contact partner and considers the different distances of the flanks through the contact ellipse. Again, there is a reference to the contact time. (7) Considering the heat flux distribution between the contacting teeth, the specific contact energy E k is given by equation (8). (8) The different heat flux leads to a change of the first equation term from γ Ek to (1-γ Ek ) in the calculation of the counterpart, so the energy input per area as a function of time for the wheel is obtained analogously. The influence of the lubricant as an intermediate medium in the contact is focused in further investigations with regard to the influence of the lubricating film thickness in a crossed helical gear flank contact. Application of the calculation to consider the contact behavior Based on gear examples of previous research work [BECK23b], calculations for a ZI, ZC and ZC-S geometry modification should be implemented to verify the calculation approach. At the same time, it should be de- ( ) = _ ( ) ( ) ( ) ( ) ( ) ( ) = + ( ) ( ) = ( ) ( ) ( ) = ( ) _ ( ) ( ) ( ) ( ) ( ) Science and Research 38 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 Considering the sliding direction φ slide1,2 , the contact body-dependent path s ellipse1,2 through the contact ellipse results, which, combined with the tangential velocity, leads to the contact time T k1,2 . The sliding path results from the product of the sliding velocity v g with the contact time T k1,2 and represents a first parameter for the friction behavior of the contacting gears. For the further local consideration of the sliding rolling contact, knowledge about the tooth friction coefficient during the contact is necessary. In the first step, an approach by Pech [PECH11] is selected for the following consideration. According to equation (3), the local tooth friction is composed of a basic friction coefficient μ 0T and a load coefficient μ 1T . These quantities depend on the sliding velocity, the resulting tooth normal force in contact and dimensional coefficients. (3) The tribological effects in the contact zone depend decisively on the maximum temperature increase. The flash temperature according to Blok is the relevant parameter according to [ISO17]. Blok [BLOK37] provides an analytical model which is able to calculate the maximum flash temperature below the heat-affected zone of the moving heat source. The temperature increase results from the diffusion of frictional heat due to contact stress under relative velocity. Local frictional energy is introduced into the contact zone and thermally dissipated from the contact zone into the substrate [WIŚN00]. The thermal depth effect largely depends on the material parameters thermal conductivity λ M , density ρ M and heat capacity c P and is described with the thermal contact coefficient B M . Due to the different material behavior of the different materials, the thermal contact coefficient must be calculated individually for both contact bodies using equation (4). (4) The theory of Blok [BLOK37] considers the tangential movement of the heat source along the contact path through the contact surface. During tooth contact, the contact force resulting from the torque is distributed to multiple teeth in contact. Locally, the contact pressure σ Hκ transmitted under relative motion v g and frictional influence μ z_contact leads to a frictional energy input. The model includes the simplification that the planar heat-affected zone is approximated with the help of a point heat source. The shape parameter A Form transforms the two-dimensional consideration into a point-shaped equivalent. As a result of the axis crossing angle, the path through the ellipse is unequal for the worm and wheel. Regarding this, the contact times T k1,2 , in which frictional energy is introduced into the surfaces, are different for the two contact partners. The effective distribution of the heat input is determined by the thermal contact coefficient _ ( ) = ( ) + ( ) = monstrated that non-involute crossed helical gear geometries offer potential in load-carrying capacity. Table 1 lists the gear parameters. Due to the different contact locations on the tooth flanks, the graphs in the following diagrams are offset. The local contact points of the start of contact A and end of contact E (on the tooth flanks) are compared in absolute terms in the parameter values. The observation is made along the parameter u in the whole contact area. Within the graphs, the steps visualize the change between the numbers of teeth in contact. The sliding velocity is similar for all three geometries due to the comparable dimensions of the gear teeth. The starting point of the optimization process was the ZI reference gear according to Pech [PECH11], which should be optimized by modifying the flank curvatures. The position and size of the contact ellipse can be influenced through design of the curvatures. Therefore the tribological potential is mainly related to the flank pressure, as shown in Figure 4-b. Figure 5 compares the different contact times of the worm and the wheel. While the worm flank is in contact for a very short time, the plastic wheel remains in contact significantly longer. In particular, the ZC contact end shows a long contact time, while the values of ZI and ZC-S are comparable. With regard to the coefficients of friction according to Figure 6-a, the difference between the individual contact points is small. While the tooth friction coefficient for ZI is constant in the contact area, the values for the ZC and ZC-S geometry decreases along the path of contact. Combined with the sliding velocity and the flank Science and Research 39 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 g y ZI ZC ZC-S Normal section Center distance a 30 mm Number of tooth z 1 / z 2 1 / 40 Normal pressure angle n 20° 5° 15° Normal modulus m n 1.25 mm Helical angle 1 / 2 82.493° / 7.507° Tip diameter d a1 / d a2 12.068 mm / 52.932 mm 9.818 mm / 55.182 mm 10.250 mm / 54.750 mm Reference diameter d 1 / d 2 9.568 mm / 50.432 mm Center diameter d m1 / d m2 9.568 mm / 50.432 mm 10 mm / 50 mm Root diameter d f1 / d f2 6.443 mm / 47.307 mm 4.193 mm / 49.557 mm 4.625 mm / 49.125 mm Tooth thickness facto s mx * 0.5 0.3 0.25 Contact ratio 1.837 2.247 1.708 Operating point n 1 / T 2 1500 1/ min -1 / 8 Nm Table 1: Parameter sizes of the ZI, ZC and ZC-S geometry modification [BECK23a] Figure 4: Comparison of the a) sliding velocity and b) flank pressure in the contact area u two thirds of the contact area, which, however, experience an exponential increase towards the contact end. The ZC-S geometry shows lower flash temperatures in the entire contact area. Science and Research 40 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 pressure in multiple tooth contact, the flash temperature of Blok is obtained according to Figure 6-b. The ZI geometry is used as the initial point for the comparison. The ZC geometry shows the lowest flash temperatures in Figure 5: Contact time of the tooth flanks along the path of contact for a) worm and b) wheel Figure 6: Path of a) tooth friction coefficient in the contact area and b) flash temperature in the tooth contact as local temperature increase according to Blok Figure 7: Change in the Hertzian contact area A K For the consideration of the specific contact energy, the area of the Hertzian contact ellipse must be determined first (Figure 7). The contact area A k represents the zone of influence in which the heat flow diffuses into the material during the contact time. The larger the area, the better the heat can be dissipated from the tooth contact. However, a balance must be found with the increasing sliding paths. From the comparison, ZI and ZC-S have similar contact areas with only minor differences along the area of contact. The ZC geometry has an advantageously large contact area at the beginning, but it becomes small at the end of contact. Combined with the high contact velocities, increased wear tendencies are expected in practice at the corresponding locations on the wheel flank. Considering the distribution of the heat flux between the worm and the wheel as well as their heat absorption capacity in the zone of influence A k during the contact time T k1,2 , the specific contact energy E k is obtained according to Figure 8. Both at the wheel and the worm, compared to ZI, the ZC geometry shows a lower energy input in the first half of contact, which is, however, dominated by an increase at the end of contact. The ZC geometry provides a noticeable improvement in contact behavior, but experience a high surface stress at the contact end due to large sliding friction. The plastic teeth have been designed to be significantly thicker than the ZI geometry and will better resist increased frictional wear. For the ZC-S geometry, the calculation results in the contact area are more favorable than for ZI and, in combination with the likewise positive development of the flash temperature θ fl , indicate a complete gear optimization. Verification of the theoretical consideration with test rig results After the theoretical calculation, practical tests of the ZI and ZC geometry are used to verify the approach. Typically, a grease lubrication is used for applications with crossed helical gears, so the test examples are lubricated with equal grease (Klübersynth LI 44-22). In a first step, it is assumed that all three geometries have a similar lubricant-influence and therefore the lubricant film analysis is outsourced to further investigations. The test rigs have speed and torque measurement in front of and behind the test gearbox. A temperature measurement is included directly under the tooth contact [BECK23b]. The previously described calculations show that there is a significant difference in the surface stress of the plastic wheel between the ZI and ZC geometry, especially at the contact end. While the coefficient of friction and the local flash temperature according to Blok are lower in a large range than with ZI, increased surface stress follows due to high sliding friction and high contact times in a small contact area at the ZC contact end. As a result of the increased flash temperature and specific contact energy at ZC, higher wear can be expected at the contact end for the same load. Additionally to the influence in wear, the heat development results in a reduction in the strength of the gear wheel material and leads to an increase in plastic deformation. The intensity depends on the plastic material and is currently taken into account in FEM simulations. The thicker ZC plastic teeth prove to be advantageous here [BECK23b] . After the test, the plastic gears were dismounted at defined time steps. The material removal is determined by measuring the surface contour with the Alicona Infinite- Focus G4 optical measuring system. Figure 9 shows the results. The operating point corresponds to Table 1. Due to the favorable contact conditions, the wear pattern at the wheel tip shows less material removal at the start of contact. Because of the worsening contact behavior at the contact end, wear increases and a visible growth of the contact surface occurs in the root area of the ZC wheel (magenta area, Figure 9). In addition, the increased surface stress leads to a roughening of the material surface. The results prove that a high sliding path at the contact end and an increased flash temperature and specific contact energy lead to higher wear at the ZC tooth root. The reduction in pressure and optimized curvature in the remaining contact area lead to less wear overall. A detailed description of the test operation is given in [BECK23b]. Additionally to wear, the operating behavior was examined. Because of the geometry optimization of a ZI geo- Science and Research 41 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 Figure 8: Specific contact energy in the contact area at the worm (a) and at the wheel (b) Summary Based on new calculation approaches for the design and consideration of general crossed helical gear geometries in the entire contact area, it is possible to analyze the sliding-rolling behavior in terms of friction energy more precisely. With a detailed breakdown of the contact behavior, it is possible to transfer the temperature approaches of ISO 6336-20 [ISO17] from spur gears to general crossed helical gears with point contact and a threedimensional contact motion. The modification relates to the Hertzian pressure taking into account multiple tooth engagement and integrates a tooth friction coefficient that changes along the contact area. The temperature increase resulting from a load-dependent frictional energy input leads to a material-dependent distribution of the Science and Research 42 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 metry to a ZC geometry, the efficiency increases by 4 % and the sump temperatures directly under the tooth contact is reduced by 4 - 5 K (Figure 10). The results show that the ZC geometry has an advantage in terms of efficiency and sump temperature although the contact behavior at the contact end is unfavorable. Based on the test results, a non-involute crossed helical gear with completely improved load-carrying capacity and a reduction of the sliding and frictional effects at the contact end is possible to calculate. The ZC-S gear is part of current research work. Like the ZC geometry, the experimental results confirms the theoretical calculation. The critical contact situation at the end of contact no longer exists. ZI after 100 h ZC after 100 h 23,3 23,7 24,1 24,5 24,9 25,3 25,7 26,1 26,5 -200 -100 0 100 200 Transverse coordinate z [mm] 23,3 23,7 24,1 24,5 24,9 25,3 25,7 26,1 26,5 -2,0 -1,6 -1,2 -0,8 -0,4 0,0 Transverse coordinate z [mm] Transverse coordinate y [mm] 24,4 24,8 25,2 25,6 26,0 26,4 26,8 27,2 27,6 -2,0 -1,6 -1,2 -0,8 -0,4 0,0 Transverse coordinate z [mm] Transverse coordinate y [mm] 24,4 24,8 25,2 25,6 26,0 26,4 26,8 27,2 27,6 -200 -100 0 100 200 Transverse coordinate z [mm] Deviation to the real profile [ m] μ Deviation to the real profile [ m] μ ZI ZC Largest wear in the middle flank area Largest wear in the root flank area Figure 9: Wear behavior and false-color images to illustrate material removal as a function of surface stress and flank geometry, the middle pictures show the growth from 1 h, 25 h, 50 h and 100 h ZI ZC - 8 Nm Legend Efficiency η ges [%] 0 50 100 150 200 250 300 10 30 50 70 90 60 80 40 20 - 8 Nm Grease sump temperature [°C] t S Ti [h] me t 0 50 100 150 200 250 300 0 10 20 30 40 25 35 15 5 Time [h] t a) b) Figure 10: a) Measured grease sump temperature shows a reduction with b) simultaneous efficiency increase of a ZI and ZC geometry under identical operating conditions [BECK23a] heat flow into the material. Due to the low thermal conductivity of the plastic, the largest proportion of heat is dissipated via the worm. Practical tests verify the calculation algorithm and the results. The critical contact behavior at the contact end, shown in the example of the ZC geometry, is practically confirmed by increased material removal in the root area of the wheel. As a result of the thicker ZC plastic teeth, the stiffness increases, so tooth bending and flank deformation is lower. To reduce the increased wear in the root, the acting load at the contact end should be reduced. This can be achieved by profile modifications. The ZC-S geometry represents an example, which leads to a complete geometry optimization. This leads to the following summary. The ZC-S geometry is particularly suitable for performance gears with explicit speed and torque requirements. The ZC geometry is suitable for applications with less requirements for wear and a sliding motion. The design of crossed helical gears depends essentially on the requirements of the applications, so the optimum geometry needs to be selected individually. References [BECK23a] Becker, L.; Tenberge, P.: Crossed helical gears - simulative studies and experimental results on non-involute geometries, In: Forschung im Ingenieurswesen (2023), H. 87 [BECK23b] Becker, L.: Erweiterte Schraubradberechnung für allgemeine Flankenformen zur Ermittlung der örtlichen Belastungen. Dissertation Ruhr-Universität Bochum, 2023 [BLOK37] Blok, H.: Theoretical study of temperature rise at surfaces of actual contact under oiliness lubricating conditions, In: Proc. General Disc. Lubrication (1937) [BOEH20] Boehme, C.: Berechnungsverfahren zur Erweiterung der Anwendungsgrenzen und der Optimierung von Schraubradgetrieben. Dissertation Ruhr- Universität Bochum, 2020 [ISO17] Norm ISO/ TS 6336-20. Calculation of load capacity of spur and helical gears - Part 20: Calculation of scuffing load - Flash temperature method, 2017 [LOOS15] Loos, J.; Kruhöffer, W.: Einfluss der Reibbeanspruchung auf die WEC-Bildung in Wälzlagern, In: Tribologie und Schmierungstechnik (2015), H. 62 [PECH11] Pech, M.: Tragfähigkeit und Zahnverformung von Schraubradgetrieben der Werkstoffpaarung Stahl/ Kunststoff. Dissertation Ruhr-Universität Bochum, 2011 [STUH22] Stuhler, P.; Nagler, N.: Stand der Technik: Anschmierungen in Radial-Zylinderrollenlagern, In: Forschung im Ingenieurwesen (2022), H. 86, S. 1-20 [SUCK13] Sucker, J.: Entwicklung eines Tragfähigkeitsberechnungsverfahrens für Schraubradgetriebe mit einer Schnecke aus Stahl und einem Rad aus Kunststoff. Dissertation Ruhr-Universität Bochum, 2013 [WIŚN00] Wiśniewski, M.: Elastohydrodynamische Schmierung: Grundlagen und Anwendungen, Renningen-Malmsheim, expert-Verl., 2000 Science and Research 43 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 DOI 10.24053/ TuS-2024-0005 News 44 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 Tribology is boring? A scientific conference can’t be fun? Not at all! In 2023, the GfT Tribology Conference proved once again how much fun tribology is and how exciting research can be. The new format of science communication is derived from so-called science slams * [1] and is named TriboSlam. The TriboSlam aims to offer conference participants the opportunity to present their tribology research or exciting tribological observations from everyday life in a short and entertaining way, without compromising the content [2]. The slammers have 360 seconds to entertain the audience with humour, fun and originality. The advantages of such a format are clear: the slammers have a good opportunity to gain experience in communicating their own research and to improve their own presentation skills [3]. While enjoying a glass of beer or a glass of wine, the audience learns about important scientific topics in an entertaining way, as the slammers have to try to inspire the audience with the knowledge that they themselves have worked hard to acquire over many hours [2]. The first round of the TriboSlam took place in 2022. From beer lubrication, preventative measures for reproduction, classic fairy tales, communication problems, daily care routines, chivalrous material interpretation, skin care risks and the creation of the “Golden Mic” award, there were many tribological highlights on offer. After the kick-off in the first year, the success story of the TriboSlam continued. Niklas Bauer, who was the winner of the first TriboSlam with his entry “Alles dicht oder was? ” ** , replaced moderators Patrick Beau, Stephan Henzler and Mirco Kröll in 2023. The following year, the slammers competed with topics ranging from shoe lacing and animal tribology to “Schmiergeld” *** and sealing technology to a tribological song. Ratings were based as in the year before on the volume of applause from the audience. In the end, the winner was Victoria Schröder with her presentation “Dichtung und Wahrheit - die Schröder’sche Hypothese” **** . Victoria van Camp and Roland Larsson from the University of Lulea took part out of competition, who - in the true tradition of the Eurovision Song Contest (Lulea Calling) or as an outside bet (Thomas Gottschalk “Wetten, dass...? ”) depending on your point of view - were switched on online from Sweden. News: Gesellschaft für Tribologie The GfT TriboSlam - be part of it this year! Authors: Niklas Bauer, Victoria Schröder, Stephan Henzler, Patrick Beau, Mirco Kröll, Mirjam Bäse News 45 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 Of course, all of this is much different from the seriousness of the scientific talks that are usually being given at conferences. But for GfT, this venture has paid off! The participants and the audience were enthusiastic about the new format of tribology communication every year. The commitment of the participants, their creativity, but also their conviction to believe in this format, made the opening evening of the conference a complete success. This means that the TriboSlam will hopefully remain in the tribological world for a very long time to come. For the organisation, special thanks go to Mirjam Bäse and Irene Kollenbrandt, whose tireless efforts behind the scenes have made the TriboSlams the success they are today. The presentations are available on the YouTube channel of the Gesellschaft für Tribologie e.V. and thus safely preserved for posterity. In this way, in addition to the tribological specialist audience, people from outside the field can also gain an exclusive and understandable insight into tribology. This way, the association contributes to reaching people who are confronted with tribological issues, but do not know that tribology offers solutions for them. Feel free to take a look at our YouTube channel. The corresponding QR codes can be found below. And if you have acquired a taste for it, then we cordially invite you to take part in this year’s 65 th Tribology Conference! Become part of the new and at the same time long standing tradition of GfT: curiosity, fun and passion for tribology! * A science slam is a short scientific presentation tournament in which scientists present their research topics to an audience within a set time limit. The focus is on the popular scientific communication of scientific content; the evaluation is carried out by the audience. In addition to the scientific content, the comprehensibility and entertainment value of the presentation are also assessed. ** A German expression with a double meaning. The phrase can mean: “Are you still in your right mind or what? ” or “Everything sealed or what? ” *** A German play on words that combines the words lubrication and money and means something like bribe money. **** In German, the same word is used for the word Poetry and the word shaft seal. Poetry/ Sealing and truth - the Schröder hypothesis. In addition, “Dichtung und Wahrheit” is the title of an autobiography of the famous German poet Johann Wolfgang von Goethe. [1] https: / / de.wikipedia.org/ wiki/ Science-Slam, visited 03.03.2024 [2] https: / / www.scienceslam.de/ was-ist-ein-science-slam/ , visited 03.03.2024 [3] https: / / www.wissenschaftskommunikation.de/ format/ science-slam/ , visited 03.03.2024 News 46 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 News: Gesellschaft für Tribologie Poster zu folgenden Themen ngsverhalten - Verschleißverhalten etervaria琀onen echnologien rung - Metalle - Polymere - Elas e - Keramik - Leichtbauwerksto昀e rungstechnik e - Kühlschmiersto昀e - innova琀ve e - Re-Ra nierung - Nachhal琀gkeit i琀ven mit Metallober昀ächen onsschichten, Mechanismen, tsto昀verarbeitung - Minimalmen earbeitung r昀ächentechnologien nd -verfahren - tribochemische noeigenscha昀en-Charakterisierung ntriebstechnik hnradkontakt - Kupplungen - Di昀e nd hydraulische Systeme ialwellendichtungen - berührungseri昀ka琀on von Prüfverfahren - Rei technik - Online Monitoring sche Methoden - Maschinelles Ler - - Neuronale Netzwerke - Use Cases Antriebe - elektrische Antriebe - endung von e-Fuels Landwirtscha昀 aterialtransport - Bodenbearbei hrzeuge - Erzmühlen n - biomedizinische Werksto昀e - p琀k - Smart Surface Technologie We invite you to submit abstracts for papers and posters for the following topics: » Tribosystems Model fric琀on systems - fric琀onal behavior - wear behavior - contact mechanics - parameter varia琀ons » Materials and Materials Technology Tribological characteriza琀on - metals - polymers - elasto mers - compounds - ceramics - lightweight materials » Lubricants and Lubrica琀on Technology Base oils - addi琀ves - greases - metal working 昀uids - inno va琀ve 昀uids - lubrica琀on systems re-re昀ning - sustainability Special Topic: Interac琀on between Addi琀ves and Metal Surfaces Machining of metals - reac琀ve layers - mechanisms - theories and models » Machining and Forming Technology Metals and polymer processing - minimum quan琀ty lubrica - 琀on - dry machining » Thin Layers and Surface Technologies Coa琀ng materials and processes, tribochemical coa琀ngs - microand nano proper琀es - coa琀ngs characteriza琀on » Machine Elements and Transmission Technology Journal and roller bearings - gearwheel contact - clutches - di昀eren琀al gears - pneuma琀c and hydraulic systems » Sealing Technology Slip ring seals - sha昀 seals - non-contact systems » Tribometry Tribological test chain - veri昀ca琀on of test methods - fric - 琀on and wear tes琀ng technology - online monitoring » Databases und Data Analysis So昀ware tools - sta琀s琀cal methods - machine learning - ar琀昀cial intelligence - neural networks - use cases » Tribology in Automo琀ve Technology Chassis - conven琀onal, electrical and hydrogen drives - applica琀ons for e-fuels » Tribology for Civil Engineering and Agriculture Materials and technologies for drill heads - material t ransport - ore mills - track vehicles - mechanical soil treatment » Biotribology, Life Science Tribosystems in living organisms - biomedical tribomateri als - medical technology - product hap琀cs - smart surface technologies Programmausschuss/ Program Commi琀ee » G. Poll, Hannover (Vorsitz) » D. Bartel, Magdeburg » M. Jungk, Wiesbaden » A. Leson, Dresden » V. Popov, Berlin » A. Rienäcker, Kassel » B. Sauer, Kaiserslautern » R. Spallek, München » C. Specht, Schweinfurt » K. Stahl, München Plenarvorträge/ Plenary Speeches 2024 kann die GfT auf ihr 65-jähriges Bestehen zurückblicken. Dazu wird der erste Plenarvortrag die Historie der Tribologie in Deutschland und die Zukun昀sperspek琀ven des Fachgebietes all gemein beleuchten. 2024 will be the 65th anniversary of GfT. Therefore, the 昀rst plen ary speech will highlight the history of tribology in Germany and the future perspec琀ve of this 昀eld in general. Weitere zugesagte Plenarvorträge/ further con昀rmed plenary speeches: » C. Gachot, TU Wien: „Perfect Fric琀on in 2D - Solid Lubrica琀on with MXenes and Transi琀on Metal Carbo Chalcogenides“ » M. Marian, Pon琀昀cia Universidad Católica de Chile: “AI think, therefore AI am a Tribologist” » M. Dienwiebel, Karlsruhe Ins琀tute of Technology: “Verständnis tribologischer Mechanismen durch Kombina琀on mul琀skaliger Experimente und Simula琀on” Neues Thema/ New Topic: Tribologie in Sommersportarten/ Tribology for Summer Sports Reibungsop琀mierung bei Sportgeräten - Sportkleidung - P昀ege- und Heilmi琀el op琀mized fric琀on for sports equipment - sports clothing - care and cure products Gesellscha Tribologie e Information und Anmeldung/ Information and registration: Vortrags- und Posteranmeldung erfolgen über die Webseite: / Registra琀on of papers and posters via website: www.g昀-ev.de/ de/ tribologie-fachtagung-2024 Veröffentlichung/ Publication Tagungsband/ Conference Proceedings Zeitschri昀en/ Journals: Tribologie und Schmierungstechnik Termine/ Deadlines: Vortragsanmeldungen/ Abstract submission ........................................... 26.04.2024 Bestä琀gung der Annahme/ Con昀rma琀on of acceptance ............................... 05.06.2024 4-zeilige Zusammenfassung für das Programmhe昀/ 4-line summary for programme booklet ............ 21.06.2024 Abgabe des Manuskripts/ Manuscript submission ...................................... 23.08.2024 Tagungsort/ venue: Tagungshotel Freizeit In Dransfelder Str. 3 D-37079 Gö ngen Tagungsgebühren/ conference fees: inkl. Tagungsunterlagen, Tagungsverp昀egung und gemeinsamem Abendessen » Nichtmitglieder/ non members € 870,- » Mitglieder/ members: GfT, DGMK € 830,- » Vortragende/ speakers € 480,- » Hochschulangehörige/ University members* € 650,- » im Ruhestand oder arbeitssuchend/ re琀red or unemployed € 250,- » Studierende/ students** € 50,- * außer Professoren bzw. Ins琀tutsleiter/ excl. Professors ** bis Master bzw. Diplom/ undergraduate Tagungssprachen ........................... Deutsch und Englisch Conference languages ..................... German and English Gesellscha昀 für Tribologie e.V. Adolf-Fischer-Str. 34 52428 Jülich Telefon: +49 2461 340 79 38 Internet: www.g昀-ev.de Einladung zur 65. Tribolog 65 th German Tri 23. - 25. S Reibung, und V Fric琀on, Gesellscha昀 Tribologie e Information und Anmeldung/ Information and registration: Vortrags- und Posteranmeldung erfolgen über die Webseite: / Registra琀on of papers and posters via website: www.g昀-ev.de/ de/ tribologie-fachtagung-2024 Veröffentlichung/ Publication Tagungsband/ Conference Proceedings Zeitschri昀en/ Journals: Tribologie und Schmierungstechnik Termine/ Deadlines: Vortragsanmeldungen/ Abstract submission ........................................... 26.04.2024 Bestä琀gung der Annahme/ Con昀rma琀on of acceptance ............................... 05.06.2024 4-zeilige Zusammenfassung für das Programmhe昀/ 4-line summary for programme booklet ............ 21.06.2024 Abgabe des Manuskripts/ Manuscript submission ...................................... 23.08.2024 Tagungsort/ venue: Tagungshotel Freizeit In Dransfelder Str. 3 D-37079 Gö ngen Tagungsgebühren/ conference fees: inkl. Tagungsunterlagen, Tagungsverp昀egung und gemeinsamem Abendessen » Nichtmitglieder/ non members € 870,- » Mitglieder/ members: GfT, DGMK € 830,- » Vortragende/ speakers € 480,- » Hochschulangehörige/ University members* € 650,- » im Ruhestand oder arbeitssuchend/ re琀red or unemployed € 250,- » Studierende/ students** € 50,- * außer Professoren bzw. Ins琀tutsleiter/ excl. Professors ** bis Master bzw. Diplom/ undergraduate Tagungssprachen ........................... Deutsch und Englisch Conference languages ..................... German and English Gesellscha昀 für Tribologie e.V. Adolf-Fischer-Str. 34 52428 Jülich Telefon: +49 2461 340 79 38 Internet: www.g昀-ev.de Einladung zur V 65. Tribolog 65 th German Tri 23. - 25. Se Reibung, und V Fric琀on, Gesellscha昀 für Tribologie e.V. Information und Anmeldung/ Information and registration: Vortrags- und Posteranmeldung erfolgen über die Webseite: / Registra琀on of papers and posters via website: www.g昀-ev.de/ de/ tribologie-fachtagung-2024 Veröffentlichung/ Publication Tagungsband/ Conference Proceedings Zeitschri昀en/ Journals: Tribologie und Schmierungstechnik Termine/ Deadlines: Vortragsanmeldungen/ Abstract submission ........................................... 26.04.2024 Bestä琀gung der Annahme/ Con昀rma琀on of acceptance ............................... 05.06.2024 4-zeilige Zusammenfassung für das Programmhe昀/ 4-line summary for programme booklet ............ 21.06.2024 Abgabe des Manuskripts/ Manuscript submission ...................................... 23.08.2024 Einladung zur Vortragsanmeldung Call for Papers 65. Tribologie-Fachtagung 65 th German Tribology Conference 23. - 25. September 2024 in Gö ngen Reibung, Schmierung und Verschleiß Fric琀on, Lubrica琀on and Wear Gesellscha昀 für Tribologie e.V. Information und Anmeldung/ Information and registration: Vortrags- und Posteranmeldung erfolgen über die Webseite: / Registra琀on of papers and posters via website: www.g昀-ev.de/ de/ tribologie-fachtagung-2024 Veröffentlichung/ Publication Tagungsband/ Conference Proceedings Zeitschri昀en/ Journals: Tribologie und Schmierungstechnik Termine/ Deadlines: Vortragsanmeldungen/ Abstract submission ........................................... 26.04.2024 Bestä琀gung der Annahme/ Con昀rma琀on of acceptance ............................... 05.06.2024 4-zeilige Zusammenfassung für das Programmhe昀/ 4-line summary for programme booklet ............ 21.06.2024 Abgabe des Manuskripts/ Manuscript submission ...................................... 23.08.2024 Einladung zur Vortragsanmeldung Call for Papers 65. Tribologie-Fachtagung 65 th German Tribology Conference 23. - 25. September 2024 in Gö ngen Reibung, Schmierung und Verschleiß Fric琀on, Lubrica琀on and Wear Gesellscha昀 für Tribologie e.V. Reibung, Schmierung und Verschleiß Friction, Lubrication an Wear News 47 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 News: Gesellschaft für Tribologie d Poster zu folgenden Themen ungsverhalten - Verschleißverhalten metervaria琀onen echnologien erung - Metalle - Polymere - Elas - 昀e - Keramik - Leichtbauwerksto昀e rungstechnik 琀e - Kühlschmiersto昀e - innova琀ve me - Re-Ra nierung - Nachhal琀gkeit di琀ven mit Metallober昀ächen 琀onsschichten, Mechanismen, ststo昀verarbeitung - Minimalmen bearbeitung r昀ächentechnologien und -verfahren - tribochemische noeigenscha昀en-Charakterisierung Antriebstechnik hnradkontakt - Kupplungen - Di昀e - und hydraulische Systeme ialwellendichtungen - berührungs- Veri昀ka琀on von Prüfverfahren - Rei stechnik - Online Monitoring sche Methoden - Maschinelles Ler z - Neuronale Netzwerke - Use Cases e Antriebe - elektrische Antriebe - endung von e-Fuels Landwirtscha昀 Materialtransport - Bodenbearbei ahrzeuge - Erzmühlen n - biomedizinische Werksto昀e - ap琀k - Smart Surface Technologie We invite you to submit abstracts for papers and posters for the following topics: » Tribosystems Model fric琀on systems - fric琀onal behavior - wear behavior - contact mechanics - parameter varia琀ons » Materials and Materials Technology Tribological characteriza琀on - metals - polymers - elasto mers - compounds - ceramics - lightweight materials » Lubricants and Lubrica琀on Technology Base oils - addi琀ves - greases - metal working 昀uids - inno va琀ve 昀uids - lubrica琀on systems re-re昀ning - sustainability Special Topic: Interac琀on between Addi琀ves and Metal Surfaces Machining of metals - reac琀ve layers - mechanisms - theories and models » Machining and Forming Technology Metals and polymer processing - minimum quan琀ty lubrica - 琀on - dry machining » Thin Layers and Surface Technologies Coa琀ng materials and processes, tribochemical coa琀ngs - microand nano proper琀es - coa琀ngs characteriza琀on » Machine Elements and Transmission Technology Journal and roller bearings - gearwheel contact - clutches - di昀eren琀al gears - pneuma琀c and hydraulic systems » Sealing Technology Slip ring seals - sha昀 seals - non-contact systems » Tribometry Tribological test chain - veri昀ca琀on of test methods - fric - 琀on and wear tes琀ng technology - online monitoring » Databases und Data Analysis So昀ware tools - sta琀s琀cal methods - machine learning - ar琀昀cial intelligence - neural networks - use cases » Tribology in Automo琀ve Technology Chassis - conven琀onal, electrical and hydrogen drives - applica琀ons for e-fuels » Tribology for Civil Engineering and Agriculture Materials and technologies for drill heads - material t ransport - ore mills - track vehicles - mechanical soil treatment » Biotribology, Life Science Tribosystems in living organisms - biomedical tribomateri als - medical technology - product hap琀cs - smart surface technologies Programmausschuss/ Program Commi琀ee » G. Poll, Hannover (Vorsitz) » D. Bartel, Magdeburg » M. Jungk, Wiesbaden » A. Leson, Dresden » V. Popov, Berlin » A. Rienäcker, Kassel » B. Sauer, Kaiserslautern » R. Spallek, München » C. Specht, Schweinfurt » K. Stahl, München Plenarvorträge/ Plenary Speeches 2024 kann die GfT auf ihr 65-jähriges Bestehen zurückblicken. Dazu wird der erste Plenarvortrag die Historie der Tribologie in Deutschland und die Zukun昀sperspek琀ven des Fachgebietes all gemein beleuchten. 2024 will be the 65th anniversary of GfT. Therefore, the 昀rst plen ary speech will highlight the history of tribology in Germany and the future perspec琀ve of this 昀eld in general. Weitere zugesagte Plenarvorträge/ further con昀rmed plenary speeches: » C. Gachot, TU Wien: „Perfect Fric琀on in 2D - Solid Lubrica琀on with MXenes and Transi琀on Metal Carbo Chalcogenides“ » M. Marian, Pon琀昀cia Universidad Católica de Chile: “AI think, therefore AI am a Tribologist” » M. Dienwiebel, Karlsruhe Ins琀tute of Technology: “Verständnis tribologischer Mechanismen durch Kombina琀on mul琀skaliger Experimente und Simula琀on” Neues Thema/ New Topic: Tribologie in Sommersportarten/ Tribology for Summer Sports Reibungsop琀mierung bei Sportgeräten - Sportkleidung - P昀ege- und Heilmi琀el op琀mized fric琀on for sports equipment - sports clothing - care and cure products Wir laden Sie ein, Vorträge und Poster zu folgenden Themen anzumelden: » Tribologische Systeme Modellreibsysteme - Reibungsverhalten - Verschleißverhalten - Kontaktmechanik - Parametervaria琀onen » Werksto昀e und Werksto echnologien Tribologische Charakterisierung - Metalle - Polymere - Elas tomere - Verbundwerksto昀e - Keramik - Leichtbauwerksto昀e » Schmiersto昀e und Schmierungstechnik Grundöle - Addi琀ve - Fe琀e - Kühlschmiersto昀e - innova琀ve Fluide - Schmierungssysteme - Re-Ra nierung - Nachhal琀gkeit Spezialthema: Wechselwirkung von Addi琀ven mit Metallober昀ächen Metallbearbeitung, Reak琀onsschichten, Mechanismen, Theorien und Modelle » Zerspanungs- und Umformtechnik Metallverarbeitung - Kunststo昀verarbeitung - Minimalmen genschmierung - Trockenbearbeitung » Dünne Schichten und Ober昀ächentechnologien Beschichtungswerksto昀e und -verfahren - tribochemische Schichten - Mikro- und Nanoeigenscha昀en-Charakterisierung » Maschinenelemente und Antriebstechnik Gleitlager - Wälzlager - Zahnradkontakt - Kupplungen - Di昀e ren琀ale - pneuma琀sche- und hydraulische Systeme » Dichtungstechnik Gleitringdichtungen - Radialwellendichtungen - berührungslose Dichtungen » Tribometrie Tribologische Prü昀e琀e - Veri昀ka琀on von Prüfverfahren - Rei bungs- und Verschleißmesstechnik - Online Monitoring » Datenbanken und Datenanalyse So昀warelösungen - sta琀s琀sche Methoden - Maschinelles Ler nen - Künstliche Intelligenz - Neuronale Netzwerke - Use Cases » Tribologie in der Fahrzeugtechnik Fahrwerk - Konven琀onelle Antriebe - elektrische Antriebe - Wassersto昀antriebe - Anwendung von e-Fuels » Tribologie für Tie昀au und Landwirtscha昀 Bohrkop昀echnologien - Materialtransport - Bodenbearbei tungswerkzeuge - Ke琀enfahrzeuge - Erzmühlen » Biotribologie, Life Science Triboysteme in Lebewesen - biomedizinische Werksto昀e - Medizintechnik- Produkthap琀k - Smart Surface Technologie We invite you to submit abstra following topics: » Tribosystems Model fric琀on systems - fri - contact mechanics - para » Materials and Materials T Tribological characteriza琀o mers - compounds - cera » Lubricants and Lubrica琀on Base oils - addi琀ves - grea va琀ve 昀uids - lubrica琀on s Special Topic: Interac琀on between Add Machining of metals - re theories and models » Machining and Forming Te Metals and polymer proces 琀on - dry machining » Thin Layers and Surface Te Coa琀ng materials and pro microand nano proper琀e » Machine Elements and Tra Journal and roller bearings di昀eren琀al gears - pneum » Sealing Technology Slip ring seals - sha昀 seals » Tribometry Tribological test chain - v 琀on and wear tes琀ng tech » Databases und Data Analy So昀ware tools - sta琀s琀cal ar琀昀cial intelligence - neu » Tribology in Automo琀ve T Chassis - conven琀onal, e applica琀ons for e-fuels » Tribology for Civil Enginee Materials and technologie port - ore mills - track veh » Biotribology, Life Science Tribosystems in living orga als - medical technology - technologies Programmausschuss/ Program Commi琀ee » G. Poll, Hannover (Vorsitz) » D. Bartel, Magdeburg » M. Jungk, Wiesbaden » A. Leson, Dresden » V. Popov, Berlin » A. Rienäcker, Kassel » B. Sauer, Kaiserslautern » R. Spallek, München » C. Specht, Schweinfurt » K. Stahl, München Plenarvorträge/ Plenary Speeches 2024 kann die GfT auf ihr 65-jähriges Bestehen zurückblicken. Dazu wird der erste Plenarvortrag die Historie der Tribologie in Deutschland und die Zukun昀sperspek琀ven des Fachgebietes all gemein beleuchten. 2024 will be the 65th anniversary of GfT. Therefore, the 昀rst plen ary speech will highlight the history of tribology in Germany and the future perspec琀ve of this 昀eld in general. Weitere zugesagte Plenarvorträge/ further con昀rmed plenary speeches: » C. Gachot, TU Wien: „Perfect Fric琀on in 2D - Solid Lubrica琀on with MXenes and Transi琀on Metal Carbo Chalcogenides“ » M. Marian, Pon琀昀cia Universidad Católica de Chile: “AI think, therefore AI am a Tribologist” » M. Dienwiebel, Karlsruhe Ins琀tute of Technology: “Verständnis tribologischer Mechanismen durch Kombina琀on mul琀skaliger Experimente und Simula琀on” Neues Thema/ New Topic: Tribologie in Sommersportarten/ Tribology for Summer Sports Reibungsop琀mierung bei Sportgeräten - Sportkleidung - P昀ege- und Heilmi琀el op琀mized fric琀on for sports equipment - sports clothing - care and cure products r zu folgenden Themen rhalten - Verschleißverhalten aria琀onen - Metalle - Polymere - Elas eramik - Leichtbauwerksto昀e ühlschmiersto昀e - innova琀ve -Ra nierung - Nachhal琀gkeit mit Metallober昀ächen hichten, Mechanismen, verarbeitung - Minimalmen entechnologien verfahren - tribochemische nscha昀en-Charakterisierung kontakt - Kupplungen - Di昀e raulische Systeme lendichtungen - berührungs- 琀on von Prüfverfahren - Rei ik - Online Monitoring We invite you to submit abstracts for papers and posters for the following topics: » Tribosystems Model fric琀on systems - fric琀onal behavior - wear behavior - contact mechanics - parameter varia琀ons » Materials and Materials Technology Tribological characteriza琀on - metals - polymers - elasto mers - compounds - ceramics - lightweight materials » Lubricants and Lubrica琀on Technology Base oils - addi琀ves - greases - metal working 昀uids - inno va琀ve 昀uids - lubrica琀on systems re-re昀ning - sustainability Special Topic: Interac琀on between Addi琀ves and Metal Surfaces Machining of metals - reac琀ve layers - mechanisms - theories and models » Machining and Forming Technology Metals and polymer processing - minimum quan琀ty lubrica - 琀on - dry machining » Thin Layers and Surface Technologies Coa琀ng materials and processes, tribochemical coa琀ngs - microand nano proper琀es - coa琀ngs characteriza琀on » Machine Elements and Transmission Technology Journal and roller bearings - gearwheel contact - clutches - di昀eren琀al gears - pneuma琀c and hydraulic systems » Sealing Technology Slip ring seals - sha昀 seals - non-contact systems » Tribometry Tribological test chain - veri昀ca琀on of test methods - fric - 琀on and wear tes琀ng technology - online monitoring » Databases und Data Analysis Programmausschuss/ Program Commi琀ee » G. Poll, Hannover (Vorsitz) » D. Bartel, Magdeburg » M. Jungk, Wiesbaden » A. Leson, Dresden » V. Popov, Berlin » A. Rienäcker, Kassel » B. Sauer, Kaiserslautern » R. Spallek, München » C. Specht, Schweinfurt » K. Stahl, München Plenarvorträge/ Plenary Speeches 2024 kann die GfT auf ihr 65-jähriges Bestehen zurückblicken. Dazu wird der erste Plenarvortrag die Historie der Tribologie in Deutschland und die Zukun昀sperspek琀ven des Fachgebietes all gemein beleuchten. 2024 will be the 65th anniversary of GfT. Therefore, the 昀rst plen ary speech will highlight the history of tribology in Germany and the future perspec琀ve of this 昀eld in general. Weitere zugesagte Plenarvorträge/ further con昀rmed plenary speeches: » C. Gachot, TU Wien: „Perfect Fric琀on in 2D - Solid Lubrica琀on with MXenes and Transi琀on Metal Carbo Chalcogenides“ » M. Marian, Pon琀昀cia Universidad Católica de Chile: “AI think, therefore AI am a Tribologist” » M. Dienwiebel, Karlsruhe Ins琀tute of Technology: “Verständnis tribologischer Mechanismen durch Kombina琀on anzumelden: » Tribologische Systeme Modellreibsysteme - Reibungsverhalten - Verschleißverhalten - Kontaktmechanik - Parametervaria琀onen » Werksto昀e und Werksto echnologien Tribologische Charakterisierung - Metalle - Polymere - Elas tomere - Verbundwerksto昀e - Keramik - Leichtbauwerksto昀e » Schmiersto昀e und Schmierungstechnik Grundöle - Addi琀ve - Fe琀e - Kühlschmiersto昀e - innova琀ve Fluide - Schmierungssysteme - Re-Ra nierung - Nachhal琀gkeit Spezialthema: Wechselwirkung von Addi琀ven mit Metallober昀ächen Metallbearbeitung, Reak琀onsschichten, Mechanismen, Theorien und Modelle » Zerspanungs- und Umformtechnik Metallverarbeitung - Kunststo昀verarbeitung - Minimalmen genschmierung - Trockenbearbeitung » Dünne Schichten und Ober昀ächentechnologien Beschichtungswerksto昀e und -verfahren - tribochemische Schichten - Mikro- und Nanoeigenscha昀en-Charakterisierung » Maschinenelemente und Antriebstechnik Gleitlager - Wälzlager - Zahnradkontakt - Kupplungen - Di昀e ren琀ale - pneuma琀sche- und hydraulische Systeme » Dichtungstechnik Gleitringdichtungen - Radialwellendichtungen - berührungslose Dichtungen » Tribometrie Tribologische Prü昀e琀e - Veri昀ka琀on von Prüfverfahren - Rei bungs- und Verschleißmesstechnik - Online Monitoring » Datenbanken und Datenanalyse So昀warelösungen - sta琀s琀sche Methoden - Maschinelles Ler nen - Künstliche Intelligenz - Neuronale Netzwerke - Use Cases » Tribologie in der Fahrzeugtechnik Fahrwerk - Konven琀onelle Antriebe - elektrische Antriebe - Wassersto昀antriebe - Anwendung von e-Fuels » Tribologie für Tie昀au und Landwirtscha昀 Bohrkop昀echnologien - Materialtransport - Bodenbearbei tungswerkzeuge - Ke琀enfahrzeuge - Erzmühlen » Biotribologie, Life Science Triboysteme in Lebewesen - biomedizinische Werksto昀e - Medizintechnik- Produkthap琀k - Smart Surface Technologie following topics: » Tribosystems Model fric琀on systems - f - contact mechanics - par » Materials and Materials Tribological characteriza琀 mers - compounds - cera » Lubricants and Lubrica琀o Base oils - addi琀ves - gre va琀ve 昀uids - lubrica琀on Special Topic: Interac琀on between Ad Machining of metals - r theories and models » Machining and Forming T Metals and polymer proce 琀on - dry machining » Thin Layers and Surface T Coa琀ng materials and pr microand nano proper琀 » Machine Elements and Tr Journal and roller bearing di昀eren琀al gears - pneum » Sealing Technology Slip ring seals - sha昀 seals » Tribometry Tribological test chain - v 琀on and wear tes琀ng tech » Databases und Data Anal So昀ware tools - sta琀s琀c ar琀昀cial intelligence - neu » Tribology in Automo琀ve Chassis - conven琀onal, applica琀ons for e-fuels » Tribology for Civil Engine Materials and technologi port - ore mills - track ve » Biotribology, Life Science Tribosystems in living org als - medical technology technologies » G. Poll, Hannover (Vorsitz) » D. Bartel, Magdeburg » M. Jungk, Wiesbaden » A. Leson, Dresden » V. Popov, Berlin » A. Rienäcker, Kassel » B. Sauer, Kaiserslautern » R. Spallek, München » C. Specht, Schweinfurt » K. Stahl, München Plenarvorträge/ Plenary Speeches 2024 kann die GfT auf ihr 65-jähriges Bestehen zurückblicken. Dazu wird der erste Plenarvortrag die Historie der Tribologie in Deutschland und die Zukun昀sperspek琀ven des Fachgebietes all gemein beleuchten. 2024 will be the 65th anniversary of GfT. Therefore, the 昀rst plen ary speech will highlight the history of tribology in Germany and the future perspec琀ve of this 昀eld in general. Weitere zugesagte Plenarvorträge/ further con昀rmed plenary speeches: » C. Gachot, TU Wien: „Perfect Fric琀on in 2D - Solid Lubrica琀on with MXenes and Transi琀on Metal Carbo Chalcogenides“ » M. Marian, Pon琀昀cia Universidad Católica de Chile: “AI think, therefore AI am a Tribologist” » M. Dienwiebel, Karlsruhe Ins琀tute of Technology: “Verständnis tribologischer Mechanismen durch Kombina琀on mul琀skaliger Experimente und Simula琀on” Neues Thema/ New Topic: Tribologie in Sommersportarten/ Tribology for Summer Sports Reibungsop琀mierung bei Sportgeräten - Sportkleidung - P昀ege- und Heilmi琀el op琀mized fric琀on for sports equipment - sports clothing - care and cure products ter zu folgenden Themen erhalten - Verschleißverhalten rvaria琀onen - Metalle - Polymere - Elas - Keramik - Leichtbauwerksto昀e Kühlschmiersto昀e - innova琀ve e-Ra nierung - Nachhal琀gkeit n mit Metallober昀ächen schichten, Mechanismen, 昀verarbeitung - Minimalmen hentechnologien -verfahren - tribochemische enscha昀en-Charakterisierung ebstechnik dkontakt - Kupplungen - Di昀e ydraulische Systeme llendichtungen - berührungsa琀on von Prüfverfahren - Rei nik - Online Monitoring Methoden - Maschinelles Ler uronale Netzwerke - Use Cases riebe - elektrische Antriebe - ng von e-Fuels wirtscha昀 rialtransport - Bodenbearbei uge - Erzmühlen biomedizinische Werksto昀e - - Smart Surface Technologie We invite you to submit abstracts for papers and posters for the following topics: » Tribosystems Model fric琀on systems - fric琀onal behavior - wear behavior - contact mechanics - parameter varia琀ons » Materials and Materials Technology Tribological characteriza琀on - metals - polymers - elasto mers - compounds - ceramics - lightweight materials » Lubricants and Lubrica琀on Technology Base oils - addi琀ves - greases - metal working 昀uids - inno va琀ve 昀uids - lubrica琀on systems re-re昀ning - sustainability Special Topic: Interac琀on between Addi琀ves and Metal Surfaces Machining of metals - reac琀ve layers - mechanisms - theories and models » Machining and Forming Technology Metals and polymer processing - minimum quan琀ty lubrica - 琀on - dry machining » Thin Layers and Surface Technologies Coa琀ng materials and processes, tribochemical coa琀ngs - microand nano proper琀es - coa琀ngs characteriza琀on » Machine Elements and Transmission Technology Journal and roller bearings - gearwheel contact - clutches - di昀eren琀al gears - pneuma琀c and hydraulic systems » Sealing Technology Slip ring seals - sha昀 seals - non-contact systems » Tribometry Tribological test chain - veri昀ca琀on of test methods - fric - 琀on and wear tes琀ng technology - online monitoring » Databases und Data Analysis So昀ware tools - sta琀s琀cal methods - machine learning - ar琀昀cial intelligence - neural networks - use cases » Tribology in Automo琀ve Technology Chassis - conven琀onal, electrical and hydrogen drives - applica琀ons for e-fuels » Tribology for Civil Engineering and Agriculture Materials and technologies for drill heads - material t ransport - ore mills - track vehicles - mechanical soil treatment » Biotribology, Life Science Tribosystems in living organisms - biomedical tribomateri als - medical technology - product hap琀cs - smart surface technologies Programmausschuss/ Program Commi琀ee » G. Poll, Hannover (Vorsitz) » D. Bartel, Magdeburg » M. Jungk, Wiesbaden » A. Leson, Dresden » V. Popov, Berlin » A. Rienäcker, Kassel » B. Sauer, Kaiserslautern » R. Spallek, München » C. Specht, Schweinfurt » K. Stahl, München Plenarvorträge/ Plenary Speeches 2024 kann die GfT auf ihr 65-jähriges Bestehen zurückblicken. Dazu wird der erste Plenarvortrag die Historie der Tribologie in Deutschland und die Zukun昀sperspek琀ven des Fachgebietes all gemein beleuchten. 2024 will be the 65th anniversary of GfT. Therefore, the 昀rst plen ary speech will highlight the history of tribology in Germany and the future perspec琀ve of this 昀eld in general. Weitere zugesagte Plenarvorträge/ further con昀rmed plenary speeches: » C. Gachot, TU Wien: „Perfect Fric琀on in 2D - Solid Lubrica琀on with MXenes and Transi琀on Metal Carbo Chalcogenides“ » M. Marian, Pon琀昀cia Universidad Católica de Chile: “AI think, therefore AI am a Tribologist” » M. Dienwiebel, Karlsruhe Ins琀tute of Technology: “Verständnis tribologischer Mechanismen durch Kombina琀on mul琀skaliger Experimente und Simula琀on” Neues Thema/ New Topic: Tribologie in Sommersportarten/ Tribology for Summer Sports Reibungsop琀mierung bei Sportgeräten - Sportkleidung - P昀ege- und Heilmi琀el op琀mized fric琀on for sports equipment - sports clothing - care and cure products anzumelden: » Tribologische Systeme Modellreibsysteme - Reibungsverhalten - Verschleißverhalten - Kontaktmechanik - Parametervaria琀onen » Werksto昀e und Werksto echnologien Tribologische Charakterisierung - Metalle - Polymere - Elas tomere - Verbundwerksto昀e - Keramik - Leichtbauwerksto昀e » Schmiersto昀e und Schmierungstechnik Grundöle - Addi琀ve - Fe琀e - Kühlschmiersto昀e - innova琀ve Fluide - Schmierungssysteme - Re-Ra nierung - Nachhal琀gkeit Spezialthema: Wechselwirkung von Addi琀ven mit Metallober昀ächen Metallbearbeitung, Reak琀onsschichten, Mechanismen, Theorien und Modelle » Zerspanungs- und Umformtechnik Metallverarbeitung - Kunststo昀verarbeitung - Minimalmen genschmierung - Trockenbearbeitung » Dünne Schichten und Ober昀ächentechnologien Beschichtungswerksto昀e und -verfahren - tribochemische Schichten - Mikro- und Nanoeigenscha昀en-Charakterisierung » Maschinenelemente und Antriebstechnik Gleitlager - Wälzlager - Zahnradkontakt - Kupplungen - Di昀e ren琀ale - pneuma琀sche- und hydraulische Systeme » Dichtungstechnik Gleitringdichtungen - Radialwellendichtungen - berührungslose Dichtungen » Tribometrie Tribologische Prü昀e琀e - Veri昀ka琀on von Prüfverfahren - Rei bungs- und Verschleißmesstechnik - Online Monitoring » Datenbanken und Datenanalyse So昀warelösungen - sta琀s琀sche Methoden - Maschinelles Ler nen - Künstliche Intelligenz - Neuronale Netzwerke - Use Cases » Tribologie in der Fahrzeugtechnik Fahrwerk - Konven琀onelle Antriebe - elektrische Antriebe - Wassersto昀antriebe - Anwendung von e-Fuels » Tribologie für Tie昀au und Landwirtscha昀 Bohrkop昀echnologien - Materialtransport - Bodenbearbei tungswerkzeuge - Ke琀enfahrzeuge - Erzmühlen » Biotribologie, Life Science Triboysteme in Lebewesen - biomedizinische Werksto昀e - Medizintechnik- Produkthap琀k - Smart Surface Technologie following topics: » Tribosystems Model fric琀on systems - f - contact mechanics - par » Materials and Materials Tribological characteriza琀 mers - compounds - cera » Lubricants and Lubrica琀o Base oils - addi琀ves - gre va琀ve 昀uids - lubrica琀on Special Topic: Interac琀on between Ad Machining of metals - r theories and models » Machining and Forming T Metals and polymer proce 琀on - dry machining » Thin Layers and Surface T Coa琀ng materials and pr microand nano proper琀 » Machine Elements and Tr Journal and roller bearing di昀eren琀al gears - pneum » Sealing Technology Slip ring seals - sha昀 seals » Tribometry Tribological test chain - v 琀on and wear tes琀ng tech » Databases und Data Anal So昀ware tools - sta琀s琀c ar琀昀cial intelligence - neu » Tribology in Automo琀ve Chassis - conven琀onal, applica琀ons for e-fuels » Tribology for Civil Engine Materials and technologi port - ore mills - track ve » Biotribology, Life Science Tribosystems in living org als - medical technology technologies » G. Poll, Hannover (Vorsitz) » D. Bartel, Magdeburg » M. Jungk, Wiesbaden » A. Leson, Dresden » V. Popov, Berlin » A. Rienäcker, Kassel » B. Sauer, Kaiserslautern » R. Spallek, München » C. Specht, Schweinfurt » K. Stahl, München Plenarvorträge/ Plenary Speeches 2024 kann die GfT auf ihr 65-jähriges Bestehen zurückblicken. Dazu wird der erste Plenarvortrag die Historie der Tribologie in Deutschland und die Zukun昀sperspek琀ven des Fachgebietes all gemein beleuchten. 2024 will be the 65th anniversary of GfT. Therefore, the 昀rst plen ary speech will highlight the history of tribology in Germany and the future perspec琀ve of this 昀eld in general. Weitere zugesagte Plenarvorträge/ further con昀rmed plenary speeches: » C. Gachot, TU Wien: „Perfect Fric琀on in 2D - Solid Lubrica琀on with MXenes and Transi琀on Metal Carbo Chalcogenides“ » M. Marian, Pon琀昀cia Universidad Católica de Chile: “AI think, therefore AI am a Tribologist” » M. Dienwiebel, Karlsruhe Ins琀tute of Technology: “Verständnis tribologischer Mechanismen durch Kombina琀on mul琀skaliger Experimente und Simula琀on” Neues Thema/ New Topic: Tribologie in Sommersportarten/ Tribology for Summer Sports Reibungsop琀mierung bei Sportgeräten - Sportkleidung - P昀ege- und Heilmi琀el op琀mized fric琀on for sports equipment - sports clothing - care and cure products Science and Research 48 Tribologie + Schmierungstechnik · volume 71 · issue 1/ 2024 Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikationswiss chaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwiss chaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Tourismus \ VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ Altphilol Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikatio issenschaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Spra issenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Tourismus \ VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ hilologie \ Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwissenschaften \ Soziologie \ Theaterwissenschaft \ Linguistik \ Literaturgeschichte \ Anglistik \ Bauwese remdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwissenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtschaft \ Touris VWL \ Maschinenbau \ Politikwissenschaft \ Elektrotechnik \ Mathematik & Statistik \ Management \ Altphilologie \ Sport \ Gesundheit \ Romanistik \ Theologie \ Kulturwiss chaften \ Soziologie \ Theaterwissenschaft \ Geschichte \ Spracherwerb \ Philosophie \ Medien- und Kommunikationswissenschaft \ Linguistik \ Literaturgeschichte \ Anglisti auwesen \ Fremdsprachendidaktik \ DaF \ Germanistik \ Literaturwissenschaft \ Rechtswissenschaft \ Historische Sprachwissenschaft \ Slawistik \ Skandinavistik \ BWL \ Wirtsc BUCHTIPP Markus Grebe Tribometrie Anwendungsnahe tribologische Prüftechnik als Mittel zur erfolgreichen Produktentwicklung Tribologie - Schmierung, Reibung, Verschleiß 1. Auflage 2021, 252 Seiten €[D] 49,90 ISBN 978-3-8169-3521-6 eISBN 978-3-8169-8521-1 expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 \ 72070 Tübingen \ Germany Tel. +49 (0)7071 97 97 0 \ Fax +49 (0)7071 97 97 11 \ info@narr.de \ www.narr.de Dieses Buch soll den interessierten Lesern aufzeigen, welche Potenziale in der anwendungsnahen tribologischen Prüftechnik (Tribometrie) stecken. Basierend auf der tribologischen Systemanalyse und der darauf aufbauenden Prüfstrategie können durch den Einsatz sinnvoller Laborprüfungen die Potenziale verschiedener Optimierungsansätze in einem sowohl zeitals auch kostentechnisch akzeptablen Rahmen gefunden werden. Im Buch wird der Unterschied zwischen einfacher Modellprüftechnik (z. B. Stift-/ Scheibe-Tests) und speziell geplanten Simulationsprüfungen auf Tribometern erläutert. Es wird aufgezeigt, wie ein anwendungsnaher Tribometerversuch und eine sinnvolle tribologische Prüfkette aufbauend auf der Systemanalyse entwickelt werden können und was dabei zu beachten ist. Dr. Markus Grebe ist seit mehr als 28 Jahren in der Tribologie tätig. Am Kompetenzzentrum für Tribologie an der Hochschule Mannheim ist er wissenschaftlicher Leiter, Laborleiter und Vorsitzender des Lenkungskreises des KTM. In dieser Funktion ist er verantwortlich für ein Team von ca. 20 technischen und wissenschaftlichen Mitarbeitern, mehr als 50 Spezialprüfstände und die dazugehörige Mikroskopie und Analytik. Er ist Mitglied in zahlreichen DIN-Arbeitskreisen, im technisch-wissenschaftlichen Beirat der Gesellschaft für Tribologie (GfT) sowie Obmann des DVM-Arbeitskreises „Zuverlässigkeit tribologischer Systeme“. Sein Fachwissen gibt er unter anderem in mehreren Fachseminaren der Forschungsvereinigung Antriebstechnik (FVA), der Deutschen Gesellschaft für Tribologie (GfT) und der Technischen Akademie Esslingen (TAE) weiter. Checklist Author information Corresponding author: F Mailing address F Telephone and fax number F eMail All authors: F Academic titles F Full name F ORCID (optional) F Research instititute / company F Location and zip code Length F Approximately: 3,500 words Data F Word and pdf documents (both with images + captions) F Additionally, please send images as tif or jpg / 300 dpi / Please send vector data as eps Manuscript F Short and concise title F Keywords: 6-8 terms F Abstract (100 words) F Numbered pictures/ diagrams/ tables (please refer to the numbers in your text) F List of works cited After the typesetting is completed, you will receive the proofs, which you are requested to review and then give your approval to start the printing process. Changes to the manuscript are no longer possible at this stage. Please also consider The editors and the publisher assume that the authors are authorized to publish all data used, that the provided texts and all visual material (images/ pictures/ illustrations) do not violate any (copy)rights of a third party, and that, where necessary, source references are provided for visual material. In cases of doubt, please obtain a printing permission from the copyright holder. Editors and publisher cannot assume liability for potential copyright infringements. Open Access Free access to knowledge is important to us. That is why you also have the opportunity to make your contribution immediately available digitally to all interested parties. This not only benefits you with an increased reach, but also researchers worldwide. In order to guarantee the high quality and substantial indexing, we are unfortunately unable to offer this service free of charge. You can obtain the full open access service for a one-off article processing charge of € 1,850.00 (plus VAT). Editor in chief Dr. Manfred Jungk eMail: manfred.jungk@mj-tribology.com Publisher expert verlag Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 D-72070 Tübingen Phone: +49 (0)7071 97 556 0 eMail: info@verlag.expert www.expertverlag.de Editor Patrick Sorg eMail: sorg@verlag.expert Phone: +49 (0)7071 97 556 57 Tribologie und Schmierungstechnik Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie | An Official Journal of Österreichische Tribologische Gesellschaft | An Official Journal of Swiss Tribology We’re looking forward to your contribution! ISSN 0724-3472 Science and Research www.expertverlag.de Zita Tappeiner, Achill Holzer, Katharina Schmitz Experimental development and validation of tribological run-in strategies to reduce friction in hydraulic applications Jan Euler, Georg Jacobs, Timm Jakobs and Julian Röder Feasibility study on preloaded flexible conical plain bearings as wind turbine main bearings Magnus Schadomsky, Johann Rauhaus, Lars Blumenthal, Detmar Zimmer, Balázs Magyar Investigation of the friction and deformation behaviour of high-speed brakes Linda Becker, Peter Tenberge Local considerations and experimental results on the contact behavior of crossed helical gears with general flank geometries