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Tribologie und Schmierungstechnik EDITOR IN CHIEF MANFRED JUNGK 3 _ 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 3 | 2024 Volume 71 Editor in chief: Dr. Manfred Jungk Tel.: +49 (0)6722 500836 eMail: manfred.jungk@mj-tribology.com www.mj-tribology.com Editorial director: Ulrich Sandten-Ma Tel.: +49 (0)7071 97 556 56 / eMail: sandten@verlag.expert Editor: Patrick Sorg Tel.: +49 (0)7071 97 556 57 / eMail: sorg@verlag.expert Dr. rer. nat. Erich Santner Tel.: +49 (0)2289 616136 / eMail: esantner@arcor.de Contributions marked with the author’s initials or full name represent the author’s opinion, not necessarily that of the editorial office. We take no responsibility for unsolicited contributions. The author is responsible for obtaining the rights to pictures. 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ISSN 0724-3472 ISBN 978-3-381-11601-0 Imprint Tribologie und Schmierungstechnik Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie | An Official Journal of Österreichische Tribologische Gesellschaft | An Official Journal of Swiss Tribology Editorial 1 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0011 To achieve greenhouse gas neutrality, it is necessary to decarbonize the energy and raw material supply, which is still largely based on fossil fuels. According to the German National Hydrogen Strategy, which began in June 2020 and progressed in April 2022, hydrogen is to make a major contribution, and the federal government wants to use it to promote the use of climate-friendly hydrogen technologies. Classically, hydrogen was produced early on from coal gasification, but with the exploration of natural gas and crude oil, it was replaced by steam reforming. This is the most economical and most widely used method of producing hydrogen. However, the carbon dioxide balance is the same as the direct combustion of fossil fuels, unless biomass is used. So far, hydrogen has been used primarily in the chemical industry, for example to produce nitrogen fertilizers or in the cracking of hydrocarbons in oil refineries, which currently amounts to around 55 TWh annually in Germany. These applications must be converted to green hydrogen-based production as far as possible. Green hydrogen, in contrast to grey hydrogen produced using fossil fuels, is hydrogen produced using excess renewable electricity by means of electrolysis. Of course, there are other areas of application for green hydrogen in addition to the existing ones, such as the steel industry and transport. In mobility applications, it is possible to use fuels obtained through Power to X in conventional combustion engines or hydrogen directly in vehicles powered by fuel cells. This mobility is an alternative for applications where the direct use of electricity is not sensible or technically not feasible, such as in air and maritime transport. The Federal Institute for Materials Research and Testing, as the scientific and technical federal authority in the business area of the Federal Ministry for Economic Affairs and Climate Protection, is dealing in detail with the challenges of a new hydrogen economy within its “H2Safety@BAM Competence Centre for Hydrogen”. The question “TRIBOLOGY - Which lubricants do we need when using hydrogen? ” is discussed there. Hydrogen environments place demands on components such as bearings, piston rings, seals and joints, as, for example, in metallic materials, once the originally protective oxide layers have been rubbed off, they are no longer renewed. This can lead to increased wear and such fresh surfaces also encourage hydrogen to penetrate metals, significantly deteriorating their mechanical properties. To answer the initial question, one should look at the global annual figures for population, gross domestic product and energy consumption over the last 70 years. These have risen steadily with very few exceptions. In comparison, however, the consumption of lubricants has remained flat over the last few decades. This can only be explained by the innovative power of the lubricant industry in cooperation with its users, when more people are moved, and more money is made with the same amount of lubricant. This lubricant industry will also succeed in making energy consumption more sustainable in the future through better lubricant technologies, and remember Tribology is everywhere. Your editor in chief Manfred Jungk Will the lubricant fall by the wayside in the National Hydrogen Strategy? Events 2 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 Events Date Place Event ► 20.11. - 21.11.24 Wiener Neustadt, Austria ÖTG Symposium 2024 ► 21.11. - 23.11.24 Kaunas, Lithuania International Tribological Conference (BALTRIB 2024) ► 22.01. - 23.01.25 Leipzig, Germany Nextlub ► 26.04. - 29.04.25 Copenhagen, Denmark 35 th Annual General Meeting ► 13.05. - 15.05.25 Brannenburg, Germany Oildoc Conference ► 18.05. - 22.05.25 Atlanta, Georgia (USA) 79 th STLE Annual Meeting & Exhibition We look forward to your contribution! The scientific journal Tribologie und Schmierungstechnik (TuS) is one of the leading publications for tribological research in Germany, Austria and Switzerland. As the official journal of the Society for Tribology (GfT) in Germany, the Austrian Tribological Society (ÖTG) and Swiss Tribology, the issues provide information on research from industry and science, current events and developments in the specialist community. Further information on the journal and publication: https: / / elibrary.narr.digital/ xibrary/ start.xav? zeitschriftid=tus&lang=en Contents 3 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 Tribologie und Schmierungstechnik Tribology - Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie An Official Journal of Österreichische Tribologische Gesellschaft An Official Journal of Swiss Tribology Volume 71, Issue 3 October 2024 5 Jaacob Vorgerd, Mathis Steinrötter, Alexander Thomas, Manuel Oehler Efficiency of high speed spur gears with an isotropic superfinishing 14 Markus Grebe, Henrik Buse, Alexander Widmann Comparison of different standard test methods for the evaluation of greases for rolling bearings under small oscillating movements 26 Matthias Kröger, Jim Gerschler, Ringo Nepp, Christian Berndt Rotary shaft seals at high temperatures 32 Dirk-Olaf Leimann Thoughts on a standardized FE-8 Test for the assessment of the WEC-carrying capacity of lubricants in rolling bearings 40 Michael Hilden, Gerd Dornhöfer, Harald Dietl Improvements in brake fluid standardization to avoid noise & wear problems 1 Editorial Will the lubricant fall by the wayside in the National Hydrogen Strategy? 2 Events Science and Research 49 News Gesellschaft für Tribologie Österreichische Tribologische Gesellschaft Columns Preface For authors Authors of scientific contributions are requested to submit their manuscripts directly to the editor, Dr. Jungk (see inside back cover for formatting guidelines). Anzeige 4 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 BOOK RECOMMENDATION expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 \ 72070 Tübingen \ Germany \ Tel. +49 (0)7071 97 97 0 \ info@narr.de \ www.narr.de The conference provides an international exchange forum for the industry and the academia. Leading university researchers present their latest findings, and representatives of the industry inspire scientists to develop new solutions. Main Topics > Trends lubricants and additives > Automotive and transport industry > Industrial machine elements and wind turbine industry > Coatings, surfaces and underlying mechanisms > Test methodologies and measurement technologies > Digitalisation in tribology > Digital Tribological Services: i-TRIBOMAT > Sustainable lubrication Target Groups > Companies in the field of lubrication, additives and tribology > Research facilities Nicole Dörr, Carsten Gachot, Max Marian, Katharina Völkel 24th International Colloquium Tribology Industrial and Automotive Lubrication Conference Proceedings 2024 1st edition 2024, 279 p. €[D] 148,00 ISBN 978-3-381-11831-1 eISBN 978-3-381-11832-8 1 Introduction Present political efforts to achieve global climate goals require efficient propulsion systems in industrial and mobile applications [DLR20]. In addition to the environmental aspects, an optimized efficiency also offers significant potential for cost savings in energy consumption. Rising prices of fossil fuels increase the need for efficiency-optimized powertrains. The concrete implementation requires the geometric and tribological utilization of the load-carrying capacity of gears flanks. The reduction in friction due to tribologically optimized slideroll contacts is beneficial for the dissipation of power losses and maximizing load carrying capacities (e.g. scuffing and micropitting sensitively depend on friction) [SNID04]. The implementation of gears with reduced friction is thus an effective action for more sustainable propulsion technology. A short-term action for efficiency optimization to the gearbox is the use of isotropic superfinished gears. Isotropic superfinishing leads to a topography with significant improvements on the frictional properties [SJÖB16; VORG23b]. An optimization in friction has already been observed in standardized tests for the scuffing load carrying capacity as well as in analogy experiments using 2-disc tribometer [SNID04; KOLL10]. The analysis regarding gear power losses conducted in this study aims to further understand this influence of the topography and confirm it for high pitch line velocities up to v t = 95 m/ s. 2 State of the art 2.1 Isotropic superfinishing of gears In conventional manufacturing, gear flanks are hard machined after heat treatment to achieve the targeted geometry [KLOC17]. The process kinematics lead to a surface texture with transverse grooves, which has functional deficits compared to surfaces used in rolling bearings. Surface topography and tribological interactions between lubricant, material, and load significantly determine the frictional effect [PRÜL15; HANS08]. The objective of isotropic superfinishing is to smooth the surface by gently removing the outer layer without additional thermomechanical stress to influence the residual stress condition. In general, the working principle involves iteratively etching the surface and subsequently removing the reactive outer layer through relative motion with grinding grits [KÖNI20]. In the context of the topographic properties, isotropic superfinishing significantly reduces the profile roughness below the capabilities of conventional profile grinding [JOAC16]. Profile ground surfaces exhibit an anisotropic topography with transverse oriented grinding grooves due to the grit of the grinding disc and the process kinematics. These grooves are often the starting condition for surface crack formation [HERG13; LOHM16; VORG19], resulting in worse wear properties. Isotropic Science and Research 5 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0012 Efficiency of high speed spur gears with an isotropic superfinishing Jaacob Vorgerd, Mathis Steinrötter, Alexander Thomas, Manuel Oehler* submitted: 12.03.2024 accepted: 15.08.2024 (peer review) Presented at the GfT Conference 2023 Gear friction is decisive for the efficiency and power density of cylindrical gears. One method to realise the highest possible efficiency is to carry out a isotropic superfinishing. In this paper, a measuring method for evaluating the gear power losses in the regime of high pitch line velocity is presented. Using this measuring method, efficiency analyses of gears with conventional profile grinding and of variants with additional isotropic superfinishing were carried out. The experimental investigations showed a significant potential in gear friction and thus efficiency due to isotropic superfinishing. The potential was confirmed independent of the gear geometry and over the entire range of pitch line velocities investigated. Keywords gear friction, isotropic superfinishing, power loss, spur gears Abstract * Dr.-Ing. Jaacob Vorgerd, Mathis Steinrötter, M.Sc., Alexander Thomas, M.Sc., Prof. Dr.-Ing. Manuel Oehler, Lehrstuhl für Antriebstechnik Ruhr-Universität Bochum Universitätsstr. 150, D-44801 Bochum lity to transfer the model to alternative topographies is not guaranteed for all models. Topography cannot be solely represented by roughness parameters and is therefore considered through empirical influence factors [PREX90; MAYE13; LÖPE15; SCHL95]. Isotropic superfinished gears are not addressed in the state of the art, so the analysis in Figure 1 considers isotropic superfinishing solely through the reduction in profile roughness to Ra = 0.1 µm. The comparison between the gear friction models shows a significant spread, with all models indicating a tendency towards decreasing coefficients of friction due to isotropic superfinishing. However, they differ in their respective weighting. 2.3 Power loss of gears Load dependent power losses P VZP are the integral of the locally dissipated frictional power (μ R · F N · v g ) in each roll angle, Eq. 1. The meshing of gears can alternatively be described by the engagement time t_E to take account for influences of elasticity on the line of action. Addressing gear geometry, the length and position of the line of action affect the formation of load dependent power loss. [VORG21; OHLE58; WIMM06] (1) In addition to the load dependent shares, the absolute gear power loss P VZ is influenced by an load independent component, Eq. 2. The load independent power loss P VZ0 results from fluid mechanical effects of the rotating gears in an air-oil environment. The primary influence factors are the pitch line velocity v_t and rheological properties of the lubricant. Furthermore, the load independent power loss is affected by the mechanical design of the gearbox components and the explicit design of the lubrication system. [MAUZ88; GREI90] (2) Science and Research 6 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0012 superfinished surfaces do not exhibit this characteristic topography. Instead, the surfaces tend to an isotropic texture with a profile roughness Ra < 0.1 µm [SOSA17; KÖNI20; JOAC16]. The time-consuming grinding process with micrometer-sized ceramic particles eliminates the profile peaks and leaves the original profile valleys from the hard machining process [NISK05]. The isotropic superfinished surface optimizes friction and relieves the surface-near stress field. Experiments on stationary disc contacts [SNID04] indicated up to Δμ = 30 % reduced friction for isotropic superfinished surfaces compared to conventionally ground surfaces. Investigations by S OSA [SOSA17] confirmed this tribological potential for gear contacts. 2.2 Gear friction Friction primarily determines the dissipation of gear power losses [NIEM03; OHLE58]. In lubricated and case-hardened gear contacts, average coefficients of friction in the range of µ m = 0.02 … 0.15 are observed [SCHO73]. The coefficient of friction depends on the local tribological condition in the gear contacts. Gears contacts are loaded by contact pressure and act under relative tangential motion. One major influence value is the pitch line velocity which relates to the relative speeds in the gear contact. During one full gear mesh the local contacts are affected by temporarily slide-roll ratios and EHL conditions. Gear friction is thus a transient phenomenon and depends on the explicit roll angle. The local tribological load, topographical condition and the lubricant formulation have the primary influence on the occurring coefficients of friction [ISO17]. In the state of the art, various models exist for calculating gear friction [SCHL95; MICH87; ISO17; JOOP18; LÖPE15; BENE61; KLEI12]. Between the models, there are significant differences under identical operating conditions. In most cases, these models are calibrated with experiments using conventionally ground gears. The abi- Legend 0 0.02 0.04 0.06 0.08 0.10 mean CoF μ m [- ] Joop [ JOOP 18b] Benedict [ BENE 61] superfinished Löpenhaus [ LÖPE 15] Klein [ KLEI 12] ISO 6336-20 [ ISO 17a] ground Löpe Klein Joop Bene ISO Δµ = - 10 % - 12 % - 35 % - 37 % - 23 %  Geometry: FZG-C  Dyn. Viscosity: η = 25 mPas Öl  Roughness: Ra = 0.5 µm (ground) Ra = 0.1 µm (superfinished) Operating point  Load: T = 200 Nm 1 v = 80 m/ s t Figure 1: Analysis regarding the influence of topography on gear friction 3 Design of experiments 3.1 Calorimetric measurement of gear power losses The conducted experiments on gear friction in this work were conducted using a setup to measure gear power losses by the calorimetric properties. This calorimetric measurement setup was integrated into a back-to-back gear test rig for high speed applications [VORG23a]. The test rig is primarily used for load-carrying investigations of spur gears, enabling pitch line velocities up to v t = 100 m/ s due to its torsionally stiff design and hydraulic load application. In the current test rig topology, utilizing calorimetry is recommended as the measurement principle for recording gear power losses. The physical working principle is based on the assumption that the power loss is entirely dissipated as heat. When achieving steady-state thermal conditions, the absorbed heat of the lubricant Q˙ oil corresponds to the gear power loss P VZ . The absorbed heat of the lubricating oil can be determined by measuring the inlet and outlet temperatures ϑ in and ϑ out and the volumetric flow rate V˙ oil as well as evaluating the lubricants properties c p and ρ oil , Eq. 3. (3) To separate the gear power losses from additional sources of dissipated heat such as bearings and sealings, the test gears are insulated with an additional housing, Figure 2. The test gears exhibit an individual lubrication system which can be regulated in volume flow, injection pressure and oil temperature. The insulation housing is made of a temperature-stable PEEK polymer. Both shafts are supported in the external steel housing, and thus, must be led out of the insulation housing. Small ! " # $ % # & ' ( gaps between the shafts and housing plates prevent the contact between these elements. Since all sealings of the test gearbox are contactless, only the bearings act as thermal error. The insulation housing prevents convective heat exchange between the heat sources, so the bearings and the test gears are only conductively connected through the common shafts. A thermal simulation of the measurement setup indicated that the conductive heat transfer through the shafts amounts to 5 - 8 % of the absolute heat for high speed conditions. The insulating setup around the test gears allows for the evaluation of the energetic state of the circulating lubricant. Due to the separated oil systems, the change in enthalpy is representative for the gear power loss. The temperature sensors used are positioned as close as possible to the injection and the returning pipe. In the inlet of the spray bars the measurement of ϑ in is carried out with a PT100-sensor embedded in the pipe. The returning oils flows back towards the tank with ambient pressure. To measure the temperature of a representative oil volume ϑ out , a siphon-like pipe connection is employed. Another PT100-sensor is immersed in the accumulated oil level. Additionally, a flow meter for measuring the oil volume flow V˙ oil is integrated into the inlet pipe, Figure 2. The conducted experiments regarding efficiency include a series of measurements in which a combination of influencing factors (pitch line velocity, topography, lubricant temperature, volume flow and gear geometry) is tested with a load spectrum of multiple loads. Figure 3 illustrates the methodology for evaluating the gear power losses. The runtime per load level is based on the time until a steady-state temperature condition is established in the test gearbox. The adjustment of load levels is automated without turning off the test rig. For each Science and Research 7 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0012 ϑ in ϑ out V oil P VZ tank measurement setup temperatur volume flow pressure pump oil properties oil density ρ [kg/ l] t ϑ emperatur oil [°C] heat capacity c p [kJ/ kg K] Figure 2: Schematic representation of the experimental setup and the measurement principle In the manufacturing of the test specimens, the gears were conventionally hard machined using the profile grinding process after case hardening. The hard machining resulted in a profile roughness Ra = 0.32 µm with a symmetric formation of valleys and peaks in the abbot curve. The achieved surface is characteristic for profile ground gears in the industrial gearbox sector [JOAC16]. The subsequent isotropic superfinishing led to a substantial reduction in the profile height. The resulting gear flanks exhibited profile roughness Ra < 0.10 µm. In the abbot curve, the surfaces indicate an isotropic, valley-dominated topography without significant profile heights, Figure 4. 4 Experimental results on gear efficiency and power losses 4.1 Results on gear power losses The conducted experiments on the influence of the topography confirm the hypothesis from the state of the art Science and Research 8 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0012 test combination, the data points represent the absolute power loss depending on torque. To separate the loaddependent and load-independent components, the data points are linearly regressed. The ordinate indicates the load-independent power loss. The difference between the absolute power loss and the load-independent component yields the load-dependent power loss. 3.2 Test specimens The efficiency measurements were conducted with two variants of test gears, Table 1. To minimize gear dynamics in the regime of high pitch line velocities, both variants were designed with a deep tooth form exhibiting a normal overlap of ε α = 2. Additionally, both variants are profile modified to reduce the impacts of elasticity and premature contact. In the design of variant Geom B, a positive profile shift was implemented to concentrate the main load events in the recess contact. In the experimental investigations, a synthetic ester-based oil was used. Preheating oil system on targeted temperature Evaluation of load dependent power loss (P = P - P ) VZP VZ VZ0 Preloading of power loop T = 500 Nm 2 Running up test rig Loading power loop acc. to load scheme Test duration 10 / 15 min (thermally steady state) through linear regression independent power loss P VZ0 Evalautaion of load n = n + 1 Increase load stage Ls 1 Ls 2 Ls 3 Ls n (Heating of components) 15 min / T = LS 2 1 10 min / T = LS 2 2 10 min / T = LS 2 3 10 min / T = LS 2 n torque T [Nm] 2 gear power loss [ kW] P VZ P VZ0 Regression Load scheme regression line data points confidence interval P VZP Figure 3: Test methodology to evaluate gear power losses Tabelle 1: Test gear geometry Denomination Symbol Unit Geom A Geom B Normale modulus m n mm 4.825 5.625 Number of teeth z 1 / z_ - 35 / 39 30 / 42 Active tooth width b 1 / b 2 mm 22 / 20 16 / 14 Normale pressure angle α n ° 22.5 20.0 Helix angle β ° 5 5 Profile shift x 1 / x 2 - 0.2000 / -0.2195 0.3000 / -0.2899 Overlap ε α - 2.00 2.00 Active tip circle d a1 / d a2 mm 183.93 / 247.69 187.04 / 248.17 that isotropic superfinishing reduces the absolute gear power losses. Figure 5-a shows the absolute gear power loss as for variant Geom A with pitch line velocities v t = 65 m/ s and v t = 80 m/ s. The respective linearity constants of the regression lines are smaller for the isotropic superfinishing samples compared to profile ground gears. On the other hand, the load independent power loss for both topography conditions is identical. Thus, topography significantly affects gear friction in the gear contacts, while the load independent shares remain unaffected. With increasing rotational speed, the load dependent components generally increase, as more gear mesh cycles accumulate per unit of time. This relationship is also evident in the measurement results. The experimental results on absolute gear power loss for variant Geom B are qualitatively similar, Figure 5-b). At both presented pitch line velocities v t = 40 m/ s and v t = 80 m/ s, the isotropic superfinished samples exhibit lower load dependent power losses. Comparing the two geometry variants at the operating point v t = 80 m/ s, the variation in oil volume flow affects the magnitude of the load independent power losses. The load independent share of variant Geom B is smaller due to the lower oil volume flow. Overall, the results for the investigated pitch line velocities indicate that load independent components constitute a significant portion of the absolute power losses. Science and Research 9 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0012 Legend Test conditions  Oil temperature: 100 °C  Oil volume flow: 14 l/ min ( ) Geom A 12 l/ min ( ) Geom B ground superfinished 0 1000 2000 3000 4000 torque T 2 [Nm] 0 5 10 15 gear power loss [kW] P V Z v = 80 m/ s t v = 65 m/ s t Objective  Gear power loss P VZ = P + P VZP VZ0 0 1000 2000 3000 4000 torque T 2 [Nm] 0 5 10 15 gear power loss P V Z [kW] b) Geom B v = 80 m/ s t v = 40 m/ s t a) Geom A Figure 5: Experimental results on gear power losses Figure 4: Resulting topography of the test gears friction. The efficiency of the isotropic superfinished samples are consistently higher than those of the profile ground samples across the entire parameter range. Furthermore, the measurement results show no significant influence of the load on efficiency. In both geometry and topography variants, constant efficiencies result for a given pitch line velocity, Figure 6-b) and Figure 6-d). The efficiencies of both geometry variants are in a similar range. 5 Conclusion and outlook In this work, a calorimetric measurement method was introduced, which is integrated into a high-speed gear test rig and is capable of determining the gear power loss of high-speed gears. Based on this method, experimental analyses were conducted to investigate the influence of isotropic superfinished gears on efficiency and gear power losses. The investigations cover the range of high pitch line velocities up to v t = 95 m/ s. Overall, the measurements confirmed the potential in efficiency achievable through a tribologically optimized topography. The efficiency benefits were also evident in the regime of high pitch line velocities. Isotropic superfinishing provides the opportunity to optimize the efficiency of propulsion systems from the perspective of manufacturing. Ultimately, an economic evaluation is Science and Research 10 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0012 4.2 Results on gear efficiency The preceding analysis results confirm that the load dependent power loss is dependent on gear friction and consequently attributable to the local tribological conditions in the gear contacts. Subsequently, the absolute power loss is corrected for the load independent components. For a qualitative influence analysis of torque and pitch line velocity on the load dependent power loss P VZP , these results are related to the mechanical power P mech . This ratio relates to the load dependent gear efficiency η VZ according to Eq. 4. The efficiency represents both the geometric conditions and the tribological conditions. (4) Figure 6 shows the experimental results on the influence of topography and gear geometry on the gear efficiency. For comparative purposes between both variants, the load is referenced to the Hertzian contact pressure at the nominal pitch point p H,C . Both geometry variants show a progressively increasing efficiency with increasing pitch line velocities, Figure 6-a) and Figure 6-c). Increasing the pitch line velocity leads to higher hydrodynamic velocities in the gear contacts and thus improves the EHL condition [DOWS68]. Increasing the fluid film thickness is beneficial for the resulting coefficient of ) * % ( ( +,-. Legend  Oil temperature: 100 °C  Geometry: Geom A  Oil volume flow: 14 l/ min Test conditions  Efficiency = 1 - P / P VZP mech η VZ Objective ground superfinished Test conditions  Geometry: Geom B  Oil volume flow: 12 l/ min  Oil temperature: 80 °C 0 50 100 pitch line velocity v [m/ s] t 99 99.2 99.4 99.6 99.8 100 p = 1250 MPa H,C 500 1000 1500 2000 Hertzian pressure p H C , [MPa] 99 99.2 99.4 99.6 99.8 100 v = 40 m/ s t 500 1000 1500 2000 Hertzian pressure p [MPa] H C , 99 99.2 99.4 99.6 99.8 100 0 50 100 pitch line velocity v t [m/ s] 99 99.2 99.4 99.6 99.8 100 gear efficiency V Z [% ] η p = 1000 MPa H,C v = 65 m/ s t a) Geom A b) Geom A c) Geom B d) Geom B gear efficiency V Z [% ] η gear efficiency VZ [%] η gear efficiency VZ [%] η Figure 6: Experimental results on gear efficiency necessary to determine whether the additional manufacturing costs offset potential additional expenses due to higher energy consumption. Funding This work is being carried out jointly with Rolls-Royce Germany as part of the KOVOHLG research project (funding number: 20T1912). The authors would like to thank Rolls- Royce Germany for their support during the project and for the opportunity to publish this work and the Federal Ministry of Economics and Climate Action (BMWK) for providing the financial resources. Abbreviations Latin symbols c P [J/ kgK] Spec. heat capacity n [rpm] Rotational speed t E [ms] Engagement time v t [m/ s] Pitch line velocity v g [m/ s] Sliding speed z [-] Number of teeth F N [N] Normal force F R [N] Friction force P VZ [W] Gear power loss P VZP [W] Load independent power loss P VZ0 [W] Load dependent power loss Q˙ oil [W] Absorbed heat (lubricant) T [Nm] Torque V˙ oil [l/ min] Oil volume flow Greek symbols η VZ [-] Gear efficiency μ R [-] Coefficient of friction ρ oil [kg/ m 3 ] Oil density ϑ out [°C] Outlet temperature ϑ ein [°C] Inlet temperature ϑ oil [°C] Oil temperature References [BENE61] Benedict, G. H.; Kelley, B. W.: Instantaneous Coefficients of Gear Tooth Friction, In: ASLE Trans. 4 (1961), S. 59-70 [DLR20] Deutsches Zentrum für Luft- und Raumfahrt (DLR): Zero emission aviation - emissionsfreie Luftfahrt. 2020 [DOWS68] Dowson, D.: Elastohydrodynamics, In: Proc.Inst. Mech. Eng (1968), H. 182, S. 151-167 [GREI90] Greiner, J.: Untersuchungen zur Schmierung und Kühlung einspritzgeschmierter Stirnradgetriebe. Dissertation TU Stuttgart, 1990 [HANS08] Hansen, B. D.: Scuffing Resistance of Isotropic Superfinished Precision Gears, In: Gear Solutions 6 (2008) [HERG13] Hergesell, M.: Grauflecken- und Grübchenbildung an einsatzgehärteten Zahnrädern mittlerer und kleiner Baugröße. Dissertation TU München, 2013 [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 [JOAC16] Joachim, F.; Kurz, N.: Influence of surface condition and lubricant on tooth flank capacity. International Tribology Colloquium Proceedings: 15th International Colloquium Tribology (2016) [JOOP18] Joop, M.: Die Fresstragfähigkeit von Stirnrädern bei hohen Umfangsgeschwindigkeiten bis 100 m/ s. Dissertation Ruhr-Universität Bochum, 2018 [KLEI12] Klein, M.: Zur Fresstragfähigkeit von Kegelrad- und Hypoidgetrieben. Dissertation TU München, München, 2012 [KLOC17] Klocke, F.; Brecher, D.: Zahnrad- und Getriebetechnik: Auslegung - Herstellung - Untersuchung - Simulation, 1. Aufl., Carl Hanser, 2017 [KOLL10] Koller, P.: Optimierung Flankentragfähigkeit: Steigerung der Zahnflankentragfähigkeit durch Kombination von Stahlbehandlung und Finishingprozess: FVA Forschungsvorhaben 521 I, 2010 [KÖNI20] König, J.: Steigerung der Zahnflankentragfähigkeit durch optimierte Fertigung und Schmierung. Dissertation TU München. - Dissertation TU München, 2020 [LOHM16] Lohmann, C.: Zusammenhang von Ermüdung, Rissbildung, Verschleiß und Graufleckentragfähigkeit an Stirnrädern. Dissertation Ruhr-Universität Bochum, 2016 [LÖPE15] Löpenhaus, C.: Untersuchung und Berechnung der Wälzfestigkeit im Scheiben- und Zahnflankenkontakt. Dissertation RWTH Aachen, 2015 [MAUZ88] Mauz, W.: Hydraulische Verluste von Stirnradgetrieben Bei Umfangsgeschwindigkeiten Bis 60 m/ s. Dissertation TU Stuttgart, 1988 [MAYE13] Mayer, J.: Einfluss der Oberfläche und des Schmierstoffs auf das Reibungsverhalten im EHD-Kontakt. Dissertation TU München. - Dissertation TU München, 2013 Science and Research 11 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0012 [SNID04] Snidle, R. W.; Evans, H. P.: Understanding Scuffing and Micropitting of Gears, In: Proceedings of the NATO Research and Technology Organization (RTO-MP) (2004), S. 1-18 [SOSA17] Sosa, M.: Running-in of gears - surface and efficiency transformation. Dissertation KTH Stockholm, 2017 [VORG19] Vorgerd, J.; Tenberge, P.: Tribologische Untersuchungen zur örtlichen und zeitlichen Grübchenbildung an Zahnradflanken, In: Tribologie und Schmierungstechnik (Bd. 66), 4-5, S. 83-85 [VORG21] Vorgerd, J.; Tenberge, P., et al.: Scuffing of cylindrical gears with pitch line velocities up to 100 m/ s, In: Forschung im Ingenieurwesen (2021), H. 86, S. 513-520 [VORG23a] Vorgerd, J.; Tenberge, P., et al.: Kalorimetrische Messung der Verzahnungsverlustleistung im Bereich hoher Umfangsgeschwindigkeiten, In: Tribologie und Schmierungstechnik 70 (2023), 22 - 30 [VORG23b] Vorgerd, J.: Wirkungsgrad und Fresstragfähigkeit schnelllaufender Stirnradverzahnungen mit chemisch glattgeschliffenen Oberflächen. Dissertation Ruhr-Universität Bochum, 2023 [WIMM06] Wimmer, J. A.: Lastverluste von Stirnradverzahnungen: Konstruktive Einflüsse, Wirkungsgradmaximierung, Tribologie. Dissertation TU München, 2006 Science and Research 12 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0012 [MICH87] Michaelis, K.: Die Integraltemperatur zur Beurteilung der Fresstragfähigkeit von Stirnradgetrieben. Dissertation TU München, 1987 [NIEM03] Niemann, G.; Winter, H.: Getriebe allgemein, Zahnradgetriebe - Grundlagen, Stirnradgetriebe, 2003 [NISK05] Niskanen P.W; Hansen B: Scuffing Resistance of Isotropic Superfinished Precision Gears. AGMA Technical Paper 05FTM13, 2005 [OHLE58] Ohlendorf, H.: Verlustleistung und Erwärmung von Stirnrädern. Dissertation TU München, 1958 [PREX90] Prexler, F.: Einfluss der Wälzflächenrauheit auf die Grübchenbildung vergüteter Scheiben im EHD-Kontakt. Dissertation TU München, 1990 [PRÜL15] Prüller, H.: Praxiswissen Gleitschleifen: Leitfaden für die Produktionsplanung und Prozessoptimierung, 2. Aufl, Wiesbaden, Springer Vieweg, 2015 [SCHL95] Schlenk, L.: Untersuchungen zur Fresstragfähigkeit von Großzahnrädern. Dissertation TU München, 1995 [SCHO73] Schouten, M. J. W.: Einfluss elastohydrodynamischer Schmierung auf Reibung, Verschleiss und Lebensdauer von Getrieben, 1973, Internet, 1973 [SJÖB16] Sjöberg, S.; Sosa, M., et al.: Analysis of efficiency of spur ground gears and the influence of running-in, In: Tribology International (2016), H. 93, S. 172-181 Science and Research 13 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 Further information and registration at www.tae.de/ go/ tribologie Attend our seminars, courses and conferences. Friction, wear and lubrication Lubricants and operating fluids Lubrication technology Lubricated machine elements A large part of our seminars is supported by the Ministry of Economic Affairs, Labour, and Housing of Baden-Württemberg with funds from the European Social Fund. Benefit from the ESF course funding and secure up to a 70 % subsidy on your participation fee. All information on eligibility for funding can be found at www.tae.de/ foerdermoeglichkeiten Tribology, friction, wear and lubrication Up to 70 % subsidy possible to ASTM D4170 [ASTM4170], which is also listed in the current NLGI high-performance multi-use specification (HPM) as a release test for lubricating greases (HPM-LL = Long Life and HPM-HL = High Load Carrying Capacity) [SHAH2020]. In Europe, the SNR- FEB2 test is frequently used [NFT1995], which is also required by many equipment manufacturers for the release of greases in blade bearings of wind turbines, among other things. In the case of standstill marks due to very small oscillation angles or vibrations, the Competence Center for Tribology Mannheim (in German: Kompetenzzentrum Tribologie Mannheim - KTM) has developed a special test (KTM-QSST) [GREB2012], which is now established in the industry. The pivoting angles vary in these three different standard tests in the range from +/ - 6° in the Fafnir test to +/ - 3° in the SNR- FEB2 test to +/ - 0.5° in the KTM-QSS test; the amplitude ratios here range from 5.5 (Fafnir) to 3.4 (SNR) to 0.55 (KTM-QSST). In addition, “classical” Fretting tests Science and Research 14 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 Introduction Rolling bearings that are frequently operated at small oscillating motions or that are exposed to vibrations at standstill can be damaged after a short time by special wear phenomena such as standstill marks or false brinelling [GREB2008, SCHA2016, GREB2017, SOSA2023, JANU2023]. Damage resulting from these phenomena can significantly reduce the fatigue life of bearings [TETO2023] and cause a wear-related increase of the friction moment [WAND2022], which endanger the function of aggregates. Since this is a special wear problem, the failures do not correlate with standard calculation approaches for life estimation such as ISO 281 [ISO281], since these are based on fatigue theories. It is therefore imperative to have application-oriented and meaningful laboratory tests available to check and evaluate the suitability of lubricants or rolling bearing modifications. In order to ensure that the results of the laboratory tests can also be used in practice, it is essential to simulate the boundary conditions prevailing in the tribological contact as well as possible [CZIC2020; GREB2021]. In the case of oscillating rolling bearings, it is therefore extremely important to consider the so-called amplitude ratio x/ 2b, which indicates the ratio between the motion of the rolling element (x) and the Hertzian contact half-axis (b) [HERT1881]. Depending on this ratio, suitable laboratory test methods must be used to test the lubricating grease in a practical manner for the particular application. For this purpose, there is the Fafnir test according Comparison of different standard test methods for the evaluation of greases for rolling bearings under small oscillating movements Markus Grebe, Henrik Buse, Alexander Widmann* Presented at the GfT Conference 2023 Rolling bearings that are often only operated at small oscillation angles or that are exposed to vibrations when stationary can show typical damage after only a short period of operation. This can be classic false brinelling damage, so-called standstill marks or fretting damage. For lubricant developers and lubricant users it is essential that laboratory test methods are available which allow a statement to be made on the suitability of a lubricant for the respective practical application. This publication explains the scientific basis for these special operating and test conditions and compares the test results of sample greases in these three standard rolling bearing tests and in a classic Fretting test under oscillating sliding friction. Keywords Rolling bearings, greases, false brinelling, standstill marks, Fafnir test, tribometry Abstract * Dr. Markus Grebe Komp.zentrum Tribologie Mannheim Dr. Henrik Buse Tribologie Engineering Mannheim Alexander Widmann Steinbeis Transferzentrum Tribologie Mannheim University of Applied Sciences Competence Center for Tribology Mannheim (KTM) Paul-Wittsack-Straße 10 68163 Mannheim - Germany according to ASTM D7594 [ASTM7594] were also carried out on the SRV test rig in point contact for this series of investigations. Scientific principles on rolling bearings at small oscillation angles or under vibrations Since the lubrication conditions in the tribologically stressed contact zone have a decisive influence on friction and wear, these must be considered very precisely within the so-called tribological system analysis [CZIC2020]. The amplitude ratio x/ 2b instead of the vibration angle allows a transfer of laboratory results also to bearings of other dimensions and loads [STAM2020]. The mere comparison of the vibration angles can lead to erroneous interpretations and thus to an unsuitable lubricant selection, since this is not solely responsible for the mechanical and kinematic contact as well as the lubrication conditions. Science and Research 15 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 Figure 1: Sketch illustrating the x/ 2b ratio at different test conditions (Based on [CATT1938, MIND1949, JOHN2003]) Figure 2: Typical wear mark with small X/ 2b ratio (here 0.55) Up to an x/ 2b ratio of 1, the contact zone between rolling element and raceway is not or not completely revealed (Figure 1). Parts of the Hertzian contact zone show micro-sliding, while the center experiences no movement up to a limit tangential force. In the case of an axial deep groove ball bearing, this leads to the typical ellipse-shaped markings (Figure 2). At x/ 2b ratios > 1, the contact is cyclically opened; typical oscillating rolling motions occur. The partial microslip movements occur only in the reversal points and are not the main cause of damage. Instead, the rolling motion leads to displacement of the lubricant (windshield wiper effect). If this cannot flow back fast enough into x/ 2b > 1. At amplitude ratios above 1, the angles are sufficiently large so that no surface element between the rolling element and raceway is permanently in contact. The contact point is thus completely opened cyclically. In general, reflow and thus relubrication of the contact point is thus possible in principle. Both from the point of view of contact mechanics [JOHN2003, GREB2012a] and from the point of view of lubrication [GREB2014, JANU2023], the tribological conditions thus differ from operating conditions in which the amplitude ratio is less than one. False brinelling occurs when a loaded oscillating rolling contact with slippage is insufficiently lubricated (starved lubrication). Adhesion and tribocorrosion result in a large quantity of wear particles, which further restrict the lubricant flow and thus replenishment. As a result, the wear particles act as an abrasive medium and, over time, deep depressions are formed, which are frequently observed in practice. According to the authors, however, the term false brinelling is thus firmly associated with the phenomenon of trough formation as it occurs only under these operating conditions. A detailed description of the causes as well as numerous results of lubrication tests is given in [GREB2020]. The extremely comprehensive publication by DE LA PRESILLA et al. summarizes the current state of science with regard to oscillating rolling bearings [PRES2023]. At amplitude ratios less than 1, visible raceway damage is clearly different from trough-shaped false brinelling damage. A typical mark that has not been altered by overrolling has an undamaged central stick zone and a damaged outer partial slip zone (Figure 2). According to GREBE [GREB2008], this damage is called the standstill mark. The damage starts at the ends of the main axis of the ellipse, i.e. in the areas with the largest microslip. Thereafter, damage development continues along the ellipse contour towards the ends of the minor ellipse axis. In the further course, the damaged and affected area enlarges due to secondary wear effects. Although the elliptical standstill marks caused by small oscillation angles, vibration, or mere elastic deformation may appear relatively harmless at first glance, there is significant localized surface damage caused by various wear mechanisms. The outer end of the elliptical wear marks exhibits tribochemical reactions and surface disruption. No changes are visible in the central sticking zone of the contact, where neither plastic damage nor sliding movements occur. At the boundary between the sticking zone and the micro-slip zone, micro-cracks occur due to large tangential forces introduced locally into the surface (Figure 3). This can significantly reduce the service life of the bearing, as tests on an FE8 testing machine with pre-damaged bearings have shown [TETO2023]. Therefore, the authors propose to clearly distinguish the terminologies and to use the term fretting for oscillating Science and Research 16 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 the raceway between the two reversal points, deficient lubrication phenomena occur [GREB2017, JANU2023]. Oil release, base oil viscosity and shear thinning effects (structural viscosity) play a crucial role [TETO23]. The sketch in Figure 1 illustrates the relationships using the example of the axial deep groove ball bearing 51206 and typical load collectives. In operational practice, the damage mechanisms are typically identified on the basis of the signs of wear on the bearing raceways. In the best case, this allows conclusions to be drawn about the operating conditions (e.g. whether the bearing is exposed to unknown vibrations). In the literature, the designation of wear damage is often ambiguous, as various terms such as false brinelling, fretting or fretting corrosion are used quasi arbitrarily for the observed wear characteristics. Some of these terms describe the resulting wear based on visual perception, while others refer more to the underlying damage mechanisms. However, a clear assignment of damage terminology is necessary, since the wear processes are based on different fundamental mechanisms, for which entirely different mechanical or lubrication approaches are required. The term fretting is usually used in German to describe a damage mechanism that occurs predominantly in reciprocal sliding contacts [GfT72009]. In combination with tribochemical oxidation, this is also referred to as fretting corrosion. The definition of fretting goes back to TOMLINSON, who in 1927 investigated the influence of vibrations between a spherical and a flat steel surface [TOML1927]. First mathematical descriptions of the processes in the contact zone were made by CATTANEO [CATT1937] and MINDLIN [1949]. Important experimental investigations on the subject of fretting with different vibration amplitudes were carried out by VINGSBO and SOEDERBERG in a ball-on-disc contact [VING1988]. They determined so-called “wear-maps”, which show the influence of partial slip and macroscopic slip (gross slip) on wear. All these investigations have in common that they were carried out under oscillating sliding friction. However, the rotation of the ball around an axis perpendicular to the normal of the contact, as is the case in a rolling bearing even at small angles, leads to different conditions from the point of view of contact mechanics [JOHN2007] and lubrication [GREB2012]. In oscillating rolling contacts, locally occurring micro-sliding motions (fretting) are usually accompanied by other mechanisms, and their individual contribution to the overall wear process can vary significantly depending on operating conditions [GREB2008, GREB2021]. The term false brinelling was first used by ALMEN in 1937 and originally referred to the formation of depressions in the bearing raceway, which can easily be mistaken for plastic indentations (“true brinelling”) as produced by Brinell hardness testing [ALME1937]. This trough-shaped damage occurs at amplitude ratios contacts with short stroke under pure sliding conditions, to revert to the term false brinelling damage for x/ 2b ratios > 1 and standstill marks for x/ 2b ratios < 1 for oscillating rolling contacts as they occur in rolling bearings. These terms describe the perception of the damage mark and indicate the underlying damage mechanisms. Grease samples used Four greases provided by an industrial partner were used in this extensive comparative study. The greases differ in terms of their NLGI class (3x NLGI 2; 1x NLGI 3), their soap thickener (lithium, calcium sulfonate) their oil release and their individual additivation. Unfortunately, only individual characteristic values may be published here. For the selection it was important that they are known to show significant differences in the Fafnir test. Fafnir test according to ASTM D4170 The American standard ASTM-D4170 [ASTM4170] was the first standardized test method for oscillating rolling bearings. It is based on the extensive investigations of HUDSON et al [HUDS1946] at the rolling bearing company FAFNIR and was introduced in 1982 by the American Society of Testing and Materials, now ASTM International, and is also known as the “Fafnir wear test”. In this test, special test bearings (thrust ball bearings similar to type 06x65, di = 16 mm; Da = 35.69 mm; 9 balls; D = 7.142 mm) are subjected to a normal force of 2450 N applied by springs (max. Hertzian contact pressure approx. 1.87 GPa, i.e. a Cdyn/ P ratio of approx. 7.9) or 4450 N (approx. 2.28 GPa, C dyn / P = 4.3) applied at a frequency of 30 Hz with a vibration angle of ± 6° for 22 hours. After testing, the bearings are visually inspected and the mass loss on the raceways is determined. According to [KLUE2010], mass losses below 5 mg are considered acceptable in the Fafnir test. The ASTM test is referred to as both the “false brinelling test” and the “fretting test.” However, due to the oscillating angles of ± 6°, which corresponds to an amplitude ratio of about 5.5, there is no reason from a scientific point of view to speak of a fretting test here, even if tribooxidation is triggered by deficient lubrication. Overall, the test has a relatively high scatter, which is why an alternative SRV test (high-frequency, linear-oscillating test machine; SRV = in German: Schwing-Reib-Verschleiß) was introduced in 2016 as a possible alternative (ASTM D7594), which will also be described later. Especially on the American market, data on Fafnir values can be found for almost all high-performance greases. In Europe, the test is not so widely used. Despite the known problems with comparability, the test has been included in the current NLGI grease specification for high performance greases (High-Performance Multiuse Greases with High Load Carrying Capacity - HPM-HL) [SHAH2020]. In addition to the standard test according to the norm, two further modifications of the test were carried out in this series of investigations in order to be able to evaluate certain influencing factors. In one test variant, the total grease quantity was reduced. According to the standard, the raceway grooves of the bearings are completely filled with grease and then drawn off. This results in a grease quantity of approx. 0.35 g. In addition, 0.65 g should be evenly distributed in the cage, resulting in a total grease quantity of 1.0 g. In another test variation (“test 2”), the Science and Research 17 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 Figure 3: FIB section at the transition between the adhesion and microslip zone shows a crack running into the depth [GREB2006] Figure 4: Overview of greases used with relevant data with the significantly lower oil release rate and the higher NLGI class takes second place in this test. In the case of calcium sulfonate samples 7 and 15, the grease is extremely thickened as a result of the particle input. Very deep troughs can be seen. Increasing the amount of grease from 0.35 g to 1 g, as specified in the standard, leads to a significant reduction in wear for all specimens, except for sample 1, which already gives a good result with the small amount of grease. The grease differences become much smaller with the larger grease amount. The mass losses are compared in detail in the overview diagram in Figure 13 and in the final discussion of the results. Science and Research 18 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 Fafnir rolling bearing was replaced by the SNR-FEB2 standard bearing 51206. The bearing adaptation in the Fafnir test rig were modified accordingly. The number of balls was reduced from 12 to 8 and the normal force was adjusted to give similar contact pressure ratios to those used in the ASTM D4170 standard test. In this test, too, the raceway grooves were drawn off flat (without further grease in the cage). The tests with 0.35 g grease show clear differences between the four greases (Figure 6). The lowest mass loss is found in sample 1, the lithium soap grease with the highest oil separation rate. Nevertheless, clear troughs can already be seen here. The other lithium soap grease Figure 5: KTM’s Fafnir tester with standard parameters and damaged test bearing Figure 6: Parameters and photo documentation of the test bearings after the test (top uncleaned, bottom cleaned) SNR-FEB2 test The SNR-FEB2 test and the associated test rig were developed by the French rolling bearing company Société Nouvelle de Roulements (SNR, now NTN-SNR Group). The test is described in the now withdrawn French standard NFT60-199 [NF1995]. The test is widely used in Europe and in the wind industry and is repeatedly requested as a clearance test. It reproduces small oscillation movements of the bearings with x/ 2b ratios significantly above 1 well. In the test, thrust ball bearings of type 51206 are tested in two units (left and right) at an angle of +/ -3° at 25 Hz and a normal force of 8000 N (max. contact pressure approx. 2.3 GPa, C dyn / P = 3.1) for 50 hours. In this test rig, too, the movement is initiated by eccentrics; the normal force is applied by springs. The test temperatures are Science and Research 19 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 Figure 8: Typical signs of wear and mass losses in SNR-FEB2-test Figure 7: Modern SNR-FEB2 test rig at the Mannheim Tribology Competence Center [Source: KTM] (9000 LW), 80 min (120000 LW)) to investigate the change over time. For each of these three individual tests, a new set of 4 balls is used, which are positioned at an unstressed location on the races. At the end of the overall test, each bearing ring thus shows 12 marks. Always 3 marks show the development over the number of cycles. The following microscopy pictures show the results after one minute (Figure 10) and 80 minutes (Figure 11). SRV fretting test according to ASTM D7594 Due to problems with the repeatability and comparability of the Fafnir test, at the urging of the U.S. aerospace industry, a new friction test on the SRV test rig (SRV = Oscillating Friction Wear Tribometer; high-frequency, linear-oscillating testing machine) was developed [FACI2007] and subsequently standardized (ASTM D7594, pure sliding, point contact, 100 N, amplitude 0.3 mm, 50 Hz, 80 °C [ASTM7594]). The authors would refer to this test as a “true” Fretting test, since a pure oscillating sliding motion is used here (no rolling! ) and the stroke is less than the Hertzian contact width (2b). The NLGI specification for high performance multipurpose greases (HPM-LL = Long Life and HPM-HL = High Load Carrying Capacity) lists this test as an alternative to the Fafnir test according to ASTM D4170. The SRV test rig is now so widespread that a more detailed presentation is omitted here. Science and Research 20 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 usually either room temperature or -20 °C. After the run, a visual inspection and weighing of the rolling bearing parts are also performed here. In SNR testing, values below 2 mg are considered very good; values up to 5 mg are still considered good [KLUE2010] (see Figure 8). When damage starts, it progresses and escalates very quickly, which can then lead to a relatively large scattering of results. Since it was known from preliminary tests that all four grease samples did not perform particularly well under the SNR-FEB2 conditions, the test duration was shortened to one hour (instead of 50 hours). In this way, differences that are later lost to secondary wear effects can still be readily seen. Grease sample 1 also performs best in this test. It can be seen that the grease is still reasonably fluid. In contrast to all three samples, no dry wear particles are visible here yet. Nevertheless, depressions can already be seen here as well (Figure 9). Quasi standstill test of the KTM (KTM-QSST) The KTM-QSST test was developed to map the lubricant influence in a macroscopically stationary bearing with alternating tangential forces, which are initiated due to very small oscillation motion or due to vibrations (QSST-quasi-stationary test) [GREB2008]. In the test, 8 of the 12 balls are removed. For this, three test times are run (for greases 1 min = 1500 load cycles (LW), 6 min Figure 9: Parameters and photo documentation of the test bearings in the SNR-FEB2 test with a test time of one hour Evaluation and discussion of the test results The following chart (Figure 13) compares all tests with a larger amplitude ratio (SNR-FEB2 and ASTM D4170 standards and variations). “ASTM D4170 (Industr.)” denotes the industry partner’s Fafnir results listed for comparison. The third row of bars (Test 1: D4170 0.35g) shows how a reduced amount of grease affects the wear result (0.35 g instead of 1.0 g as per standard). The fourth row of bars shows the results with the SNR standard bearing ARKL 51206 in the Fafnir tester with test conditions adapted to ASTM D4170 and also reduced grease quantity (only running grooves filled) (“Test 2”). It can be seen that sample 1 (lithium soap grease with high oil release) achieves a very good result in all five tests. The differences are surprisingly small, even though different test bearings, test methods and grease quantities are used here. The grease thus appears to be very well suited for all conditions. Sample 2 (lithium soap grease with lower oil release and NLGI 3) shows a significantly greater scatter. Only in the ASTM standard test at the KTM did it perform similarly well to grease 1. However, a higher Fafnir value was measured at the industry partner. This coincides with the poorer values in tests with a reduced amount of grease. In the classic SNR-FEB2 standard test, the sam- Science and Research 21 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 Figure 11: Exemplary wear marks in the KTM quasi-still test (amplitude ratio approx. 0.55) - test duration 80 min = 120,000 load cycles. Figure 10: Exemplary wear marks in the KTM quasi-still test (amplitude ratio approx. 0.55) - test duration 1 min = 1500 load cycles. in the Fafnir test with 0.35 g grease. The clear influence of the thickener is evident here, as the grease has an even lower oil release than sample 2. For grease samples 1 and 15, two greases with different thickener types and different oil release perform best on average. This shows that oil release seems to be important for the lithium soap greases. However, with a different thickener, other variables might be more important. Thus, it should be noted that the overall rheological properties of a lubricant affect performance under these particular operating conditions and the results cannot simply be reduced to a single measured value. For example, the two calcium sulfonate greases clearly show that it is not just oil release that is important. For the lithium soap greases, on the other hand, oil release is a critical factor, as other researchers have already demonstrated [SCHA2016, TETO2023, JANU2023]. The result is completely different for the tests with a small amplitude ratio. The clear but unfortunately subjective evaluation based on school grades (1 very good, 6 very poor) shows that grease sample 1 also has a clear advantage in this test (Figure 13). Nevertheless, the result is not particularly good. Calcium sulfonate grease 15, which performed very well, especially at larger angles, is rather unsuitable under these conditions. Samples 7 and 15 (both calcium sulfonate greases) differ only slightly. Sample 2 is completely unsuitable. The low oil release in combination with the higher NLGI class seems to be extremely unfavorable for the lithium soap grease. Whether calcium sulfonate greases in general perform relatively poorly under these conditions cannot yet be Science and Research 22 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 ple performed worst. Thus, the grease already has problems providing proper lubrication under these conditions. Smaller amounts of lubricant (only raceway grooves filled) seem to be particularly problematic. Sample 7 (a commercial calcium sulfonate grease) performs the worst in almost all tests. This grease also appears to have problems with small amounts of lubricant, as evidenced by the high mass losses in both tests on the Fafnir test rig at a gerase amount of 0.35 g. The sample also performs relatively poorly in the standardized ASTM D4170 or SNR-FEB2 tests. Sample 15 (modified calcium sulfonate grease) often shows the second best result. Only in the Fafnir test with reduced grease quantity has extreme wear been measured. The results clearly show that for tests with a relatively large amplitude ratio of >>1, the oil release, the thickener type and the grease quantity are decisive. Sample 1 performs by far best with an oil release of 4.64 % according to DIN 51817 [DIN51817] and is also the most reproducible. For the other samples, the statement is not quite so clear. Sample 2, with the same thickener but a higher NLGI-class and much lower oil release of 1.18 %, performs consistently much worse. Results at reduced grease levels are particularly poor. Sample 7, with the lowest oil release but based on calcium complex, also shows significant weakness at low grease levels. The results with larger amounts of lubricant are better, but also do not come close to the results of grease sample 1. Sample 15 shows good results except for the possible outlier Figure 12: Test results - comparison of different tests with an amplitude ratio of 3.3 (SNR-FEB2) or 5.5 (ASTM D 4170) generalized based on the two samples. Unfortunately, no studies by other researchers under similar conditions are known for this modern thickener type. The comparison with the “classic” Fretting test under oscillating sliding friction in the SRV shows the problem of the very high contact pressure at the beginning of this test due to the point contact. Two of the four grease samples could not be tested under these standard test conditions, as adhesive failure (seizure) already occurred in the run-in phase of the test (reduced normal force 50 N) for the two lithium soap greases investigated (Figure 14). The two calcium sulfonate greases, on the other hand, gave almost identical results (mean wear scar diameter 415 µm for sample 15 versus 397 µm for sample 7). The high contact pressure in point contact requires a special lubricant or additive chemistry for these initial conditions, which are ultimately of secondary importance in rolling bearings. The test should therefore only be used if high local contact pressures and pure sliding friction are to be expected in practice. Greases may be suitable for friction conditions in plain contact (e.g. for shaft-hub connections [BUSE2021, BUSE2022]) or for rolling bearings, even if they do not pass the SRV test according to ASTM D7594. Summary The aim of this in-house test series was to compare different laboratory test methods for evaluating lubricating greases for rolling bearings that only perform small oscillating movements or are only subjected to vibration loads (e.g. blade bearings in wind turbines). The Fafnir wear test according to ASTM D4170 and the SNR-FEB2 test frequently required in Europe can be mentioned here as standard procedures. In addition, an in-house test was carried out to simulate bearings at very low oscillation angles and vibrations (KTM QSST). The new NLGI grease specification for high perfor- Science and Research 23 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 Figure 13: Subjective Evaluation of the standstill markings by means of school grades (1 very good, 6 very poor) Figure 14: Friction coefficients in the SRV fretting test; blue and gray line: failure to run-in for samples 1 and 2 (both lithium soap greases); green sample 7, red sample 15 This series of tests was intended to show that a laboratory test must be selected that reflects practical conditions as closely as possible. This was clearly demonstrated using the specially formulated model greases. In particular, small x/ 2b ratios of less than 1 represent a major challenge for the lubricants used. Greases that still perform very well at x/ 2b ratios greater than 1 can be completely unsuitable here. The SRV fretting test (ASTM D7594) is not suitable for reliably predicting the grease behavior in a rolling bearing. If necessary, results from several tests must also be taken into account if different conditions may prevail in practice. For example, the conditions in the blade bearings of wind turbines are very different. During the adjusting motion, large x/ 2b ratios occur, while during standstill under vibration, small x/ 2b ratios have to be considered [STAM2020, SCHW2020]. The selected lubricant must therefore always represent a compromise. References ALME1937 J O Almen. Lubricants and False Brinelling of Ball and Roller Bearing; Journal of Mechanical Engineering, 59(6): 415-422, 1937 BUSE2022 H. Buse. Fretting wear tests on tribometers - basics, industrial relevance and test realisation. Promotion an der TU Bratislava; Material Sciences and Technology; Reg. No.: MTF-114753- 82998; 2022 BUSE2021 Buse, H.; Schueler, F.; Hodúlová, E. Planar Contact Fretting Test Method Applied to Solid Lubricants. Lubricants 2021, 9, 58. https: / / doi.org/ 10.3390/ lubricants9060058 CATT1938 C Cattaneo. Sul contato di due corpo elastici. Mat. Nat. Rend., 27, page 342 - 348, 1938. CZIC2020 Czichos, Habig. Tribologie-Handbuch - Tribometrie, Tribomaterialien, Tribotechnik, Springer Vieweg Wiesbaden, ISBN 978-3-658-29483-0; Edition 5, 2020 FACI2007 Faci, H. Fretting Test by SRV. In Proceedings of the ASTM Meeting Miami Beach, Naples, FL, USA, 20 June 2007. GFT72009 GfT-Arbeitsblatt Nr. 7: Tribologie - Verschleiß, Reibung - Definitionen, Begriffe, Prüfung; Gesellschaft für Tribologie: Jülich, Germany, 2009. GREB2008 M Grebe, P Feinle, and W Hunsicker. Einfluss verschiedener Faktoren auf die Entstehung von Stillstandsmarkierungen (False Brinelling Effekt). Tribologie und Schmierungstechnik, 55. Jahrgang, Heft 1-08, Expert-Verlag, Renningen, 2008, 55: 12, 2008. GREB2012 M Grebe. False Brinelling - Standstill marks at roller bearings. PhD thesis, Slovak University of Technology, Bratislava, 2012 Science and Research 24 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 mance multi-purpose greases (HPM) also requires the SRV test according to ASTM D7594 for greases with higher loads, so this was also included in the test series. Modifications to the tests were added to the test series to investigate the influence of grease amount and bearing type. For the test series, four mineral oil-based grease samples were prepared with two types of soap thickeners (lithium soap and calcium sulfonate), which gave different results in the Fafnir test. The aim of the test series was not to find a particularly suitable grease for the application, but to show that the greases give different results depending on the tribological stress collective. In the first part of the publication, the scientific principles of rolling bearings operating only at small angles or subjected to vibration were explained. In these investigations, it is important to consider the amplitude ratio x/ 2b, i.e. the ratio between the displacement of the rolling element (x) and twice the Hertzian contact width (2b). For a ratio smaller than 1, parts of the contact are never opened, which makes the re-entry of lubricant much more difficult or even impossible. At an amplitude ratio greater than 1, replenishment is possible in principle and depends strongly on the rheology of the lubricating grease. In the second part of the publication, the operating and test conditions of common and partly standardized rolling bearing test methods were presented. Four grease samples with known formulations were compared in these three standard rolling bearing tests, two modified tests and a classic Fretting test under oscillating sliding friction. It can be seen that the performance of the lubricants is highly dependent on the test conditions and that, to date, there appears to be no universal lubricant for these different operating conditions. A high oil release is advantageous for rolling bearings operating with relatively small vibration angles. But a change of the thickener can also be promising. It is important to note that greases with a high oil release rate may not achieve the desired grease operating life, which makes early relubrication necessary. A comparison with a “classical” Fretting test under oscillating sliding friction in the SRV test rig shows the problem of the very high contact pressure at the beginning of this test. Two of the four grease samples could not be tested under these standard test conditions. The high contact pressure in point contact requires a special lubricant or additive chemistry to prevent seizure, which is ultimately of secondary importance in rolling bearings. The test should therefore only be used if high local contact pressures and pure sliding friction are to be expected in practice. GREB2012a M Grebe, P Feinle, W Hunsicker. Kontaktmechanik Kontaktmechanische Beschreibung von False-Brinelling (Stillstandsmarkierungen bei Wälzlagern); Jahrestagung der Gesellschaft für Tribologie, (GfT) Tagungsband, 2010 GREB2014 M Grebe; P Blaškovitš, P Feinle. Failure of roller bearings without macroscopic motion - Influence of the pivoting angle on the contact mechanics and the wear mechanisms in the contact between roller and raceway 19th International Colloquium Tribology: Industrial and Automotive Lubrication, Esslingen, 2014; Proceedings on DVD GREB2017 M Grebe. False-Brinelling und Stillstandsmarkierungen bei Wälzlagern - Schäden bei Vibrationsbelastung oder kleinen Schwenkwinkeln. Narr Francke Attempto Verlag, ISBN: 9783825251604; DOI: 10.36198/ 9783838551609. 2017 GREB2018 M Grebe, J Molter, F Schwack, and G Poll. Damage mechanisms in pivoting rolling bearings and their differentiation and simulation. Bearing World Journal, 3: 72-85, 2018 GREB2021 M Grebe. Tribometrie - Anwendungsnahe tribologische Prüftechnik als Mittel zur erfolgreichen Produktentwicklung; 2021; 252 S.; expert verlag. ISBN 978-3-8169-3521-6 HERT1881 Hertz, H. Über die Berührung fester elastischer Körper. J. Reine Angewandte Mathematik 1881, 92, 156-171. HUDS1946 Hudson, T. Morton: Friction Oxidation; Fafnir Bearing Co.: New Britain, CT, USA, 1946. JANU2023 Januszewski, R.; Brizmer, V.; Kadiric, A. Effect of Lubricant: Properties and Contact Conditions on False Brinelling Damage. Tribol. Trans. 2023, 66, 350-363. https: / / doi.org/ 10.1080/ 10402004.2023.2183915. JOHN2003 Johnson, K.L. Contact Mechanics, 9th ed.; Cambridge University Press: Cambridge, UK, 2003; ISBN 0-521-34796-3. KLUE2010 Klueber brochure: Always with the breeze - maximum yield with optimised lubrication concepts for wind power plants, 2010 KUHN2015 Kuhn, M. Spezialschmierstoffe für False Brinell Beanspruchte Wälzlager; 11. VDI-Fachtagung “Gleit- und Wälzlagerungen 2015”; VDI-Berichte 2257, Düsseldorf, Germany; 2015; ISBN 978-3-18-092257-7. TOML1927 G A Tomlinson. The rusting of steel surfaces in contact. Proceedings of the Royal Society of London. Series A, Containing Papers of a Mathematical and Physical Character, pages 472- 483, 1927. MIND1949 R D Mindlin. Compliance of elastic bodies in contact. Journal of Applied Mechanics, pages 259-268, 1949. PRES2023 R. de la Presilla; S. Wandel; M. Stammler; M. Grebe; G. Poll; S. Glavatskih. Oscillating Rolling Element Bearing Tribology: a Review of Testing and Analysis Approaches; Tribology International, Elsevier; https: / / doi.org/ 10.1016/ j.triboint.2023.108805; 2023. VING1988 O Vingsbo and S Söderberg. On fretting maps. Wear, 1988(126): 131-147, 1988. SCHA2016 Schadow, C. Stillstehende Fettgeschmierte Wälzlager Unter Dynamischer Beanspruchung. Ph.D. Thesis, Otto-von-Guericke-Universität: Magdeburg, Germany, 2016. SHAH2020 R Shah, J Jinag, and J Kapernik. Next-generation NLGI grease specifications. NLGI Spokesman, 83(4): 63-73, 2020. SOSA2023 Sosa, Y. False brinelling: An increasing type of a rolling bearing wear Industry experts discuss how to identify false brinelling, prevention scenarios and lubricating strategies. Tribol. Lubr. Technol. 2023, 79, 60-68. STAM2020 M Stammler. Endurance test strategies for pitch bearings of wind turbines. PhD thesis, Stuttgart: Fraunhofer Verlag, 2020. TETO2022 S Tetora, C Schadow, and D Bartel. Abschlussbericht Forschungsvorhaben Nr. 540 III: Stillstehende fettgeschmierte Wälzlager unter dynamischer Belastung. FVA-Informationsblatt, 2022, Arbeitskreis Schmierstoffe und Tribologie, 1500: 4, 2022. TETO2023 Tetora, S.; Schadow, C.; Bartel, D. Influence of Grease Properties on False Brinelling Damage of Rolling Bearings. Lubricants 2023, 11, 279. https: / / doi.org/ 10.3390/ lubricants11070279 SCHW2020 F Schwack. Untersuchungen zum Betriebsverhalten oszillierender Wälzlager am Beispiel von Rotorblattlagern in Windenergieanlagen. PhD thesis, Leibniz Universität Hannover, Hannover, 2020. WAND2022 Wandel, S.; Bader, N.; Schwack, F.; Glodowski, J.; Lehnhardt, B.; Poll, G. Starvation and relubrication mechanisms in grease lubricated oscillating bearings. Article in Tribol. Int. 2022, 165, 107276. Standards ASTM4170 ASTM D4170; Standard Test Method for Fretting Wear Protection by Lubricating Greases. American Society for Testing and Materials, Pennsylvania, USA, 2016. ASTM7594 ASTM D7594; Standard Test Method for Determining Fretting Wear Resistance of Lubricating Greases under High Hertzian Contact Pressures Using a High-Frequency. Linear-Oscillation (SRV) Test Machine; American Society for Testing and Materials; Pennsylvania, USA, 2019. DIN51817 DIN 51817: 2014-08; Prüfung von Schmierstoffen - Bestimmung der Ölabscheidung aus Schmierfetten unter statischen Bedingungen (Testing of Lubricants - Determination of Oil Separation from Greases under Static Conditions). Beuth-Verlag, Berlin, Germany; 2014. ISO281 ISO 281: 2007; ISO 281 - DIN ISO: Wälzlager - Dynamische Tragzahlen und nominelle Lebensdauer. Beuth: Berlin, Germany, 2010. NFT1995 Normalisation Francaise. Nft 60-199 - aptitude à résister au faux effet brinell, 1995. Science and Research 25 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0013 velocity limits for various seals under standard conditions as a function of the shaft diameter and the seal material, e.g. [1]. However, the temperature at the seal depends very much on the heat dissipation possibilities via the shaft and, if possible, via the oil bath. If there are additional heat sources such as bearings near to the sealing area, these also has to be considered. Müller [2] provides an initial estimate of the temperature increase Δϑ at the seal contact as a function of the circumferential velocity v by (1) Leichner [3] used this approach to estimate temperature increases in the seal contact of approx. 15 K at 1500 rpm and 30 K at 3000 rpm for a shaft diameter of 78 mm. With these contact temperatures he simulates the temperature distributions in the seal using a thermal FE model. The results were temperature distributions with circular ring-like isotherms around the contact in the cross-section of the seal. In a stationary state, the heat generated in contact is divided into a heat flow through the shaft and a heat flow through the seal. However, due to the filigree cross-section of the seal and the poor thermal conductivity of elastomers, only a small proportion of the heat flow passes through the seal. Engelke [4] has assumed the distribution in relation to the thermal conductivity ratio. The thermal conductivity coefficient of steel is approx. 40 W/ m/ K and that of elastomers approx. 0.2 W/ m/ K. y = 2.5 Ks m . Science and Research 26 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0014 1 Introduction Sealing applications are critical if the seals are used at high temperatures. High velocities in the seal contact cause high friction losses and lead to a localised temperature rise on the seal contact and in the lubricating film. Therefore, dynamic shaft seals in particular quickly reach the limits of their operating conditions. If there are also high ambient temperatures, the operating conditions for the sealing application are even more difficult. This can be caused by increased oil bath temperatures or increased ambient temperatures as well as very warm components in the immediate vicinity of the seal, e.g. combustion engines, exhaust systems or process engineering reactors. Use cases under grease lubrication or minimum oil lubrication are also critical due to the lack of heat dissipation through an oil bath. This results in a strong selfheating of the seal contact at high velocities. Consequences may be ageing of the seal or the lubricant or the formation of oil carbon. The aim of this publication is to present test options for shaft seals under high temperatures and to discuss their results. 2 State of the art The temperature at the seal is often a limiting factor for sealing applications. Increased temperatures lead to faster ageing of the seal material. The very frequently used sealing material acrylonitrile butadiene rubber (NBR) is particularly sensitive to high temperatures. Therefore, seals made from hydrogenated acrylonitrile butadiene rubber (HNBR) for raised temperatures and seals made of fluororubber (FKM) for high temperatures can be used. The application limits of the seals depend on the temperature and the duration of exposure resp. the temperature profile. An initial indication is provided by diagrams from the seal manufacturers, which specify Rotary shaft seals at high temperatures Matthias Kröger, Jim Gerschler, Ringo Nepp, Christian Berndt* Presented at the GfT Conference 2024 In many applications, dynamic seals are used at high temperatures which can lead to many problems like damage or ageing of elastomer and lubricant or in formation of oil carbon. A recently built seal test rig enables very high temperatures up to above 250 °C with grease lubrication. Experiments show the influence of high temperatures on function and friction of seals. Keywords Dynamic shaft seals, high temperature, friction losses, oil and grease lubrication, test rig, ageing Abstract * Prof. Dr.-Ing. Matthias Kröger M. Sc. Jim Gerschler Dr.-Ing. Ringo Nepp Dr.-Ing. Christian Berndt Institute of Machine Elements, Design and Manufacturing Technische Universität Bergakademie Freiberg Agricolastr. 1, 09599 Freiberg According to Engelke, only approx. 0.5 % of the heat flow is transferred via the seal and over 99 % via the shaft. Engelke [4] and later Ottink [5] also calculate the temperature rise at the seal contact. It amounts to between 10 K and 17 K per 1 W friction power and per 1 mm 2 contact area. Investigations by Berndt [6] with minimum oil lubrication without an oil bath showed a heating of 2.6 K/ W depending on the friction power of a radial shaft seal without spring with a diameter of 45 mm for his test setup. Estimating a contact area of approx. 13 mm 2 , Berndt’s seal contacts heat up two to three times more than Engelke’s resp. Ottink’s seal contact. In particular, this is due to the thermodynamic boundary conditions of the different test set-ups. The friction of seals primarily consists of fluid friction and hysteresis friction. At very low velocities or start-up processes, adhesive friction also has to be taken into account, but this isn’t considered here, cp. [7, 8]. By increasing the temperature in the contact, the most important parameters for fluid friction and hysteresis friction are strongly shifted. As the temperature increases, the dynamic viscosity η of the lubricant decreases by powers of ten, see Figure 1 left. This makes it more difficult to build up the lubricating film and results in smaller lubricating film thickness h resp. the lubricating film thickness in relation to the standard deviation of roughness, see Figure 1 centre. This results in higher shear rates in the lubricating film. Assuming a Newtonian fluid, the shear stress τ in the fluid film for a circumferential velocity v is calculated as (2) = . It can be seen that a decrease in viscosity η directly reduces the shear stress in the fluid and thus the fluid friction torque. However, the decrease in lubricating film thickness increases the shear stress. The extremely strong decrease in viscosity dominates at the rotary shaft seal, so that the frictional torque decreases with increasing temperature T, see Figure 1 on the right. The modelling assumptions on which the results are based can be found in dissertation of Berndt [6]. It should be noted that various constant temperatures were specified for this simulation. In real operation, the frictional torque of the seal leads to increasing frictional power and higher temperatures as the velocity increases. Hysteresis friction is also influenced by the temperature, as the viscoelastic properties of the seal change. Hysteresis friction is caused by local cyclic deformation of the seal at the roughness of shaft and the damping effect of the viscoelastic elastomer material. The relevant frequency range at the increased temperature can be converted into a frequency range at reference temperature by means of the temperature-frequency equivalence using the WLF transformation of Williams, Landel and Ferry [9], see Figure 2 left. The figure shows the lower cut-off frequency (1 and 2) resp. the upper cut-off frequency (3 and 4) for 20 °C in blue (2 and 4) and 100 °C in red (1 and 3). For the analysed ground counterface, the hysteresis friction decreases with increasing temperature, see Figure 2 on the right. When interpreting the hysteresis friction results, it should be noted that the simulation of hysteresis friction was calculated without lubrication. This means that there is no separating lubricating film considered which reduces hysteresis friction at high velocities. It can be seen that both fluid friction and hysteresis friction decrease with temperature. This does not apply to Science and Research 27 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0014 Friction coefficient [-] Velocity [m/ s] Oil temperature [°C] Oil viscosity [Pa s] Velocity [m/ s] Rel. film thickness [-] Figure 1: Influence of temperature on viscosity of 4 lubricating oils (ISO VG 15, ISO VG 32, ISO VG 100, ISO VG 460) (left), simulated influence of contact temperature on relative lubricating film thickness (centre) and fluid friction coefficient (right) using the example of a radial shaft seal without spring for the lubricating oil ISO VG 100 In the RWDR seal test rig, the frictional torque of the seal is measured using a torque sensor which is connected to the seal via seal support and a thermal insulation to reduce the temperature at the sensor. As the seal support represents the housing and is motionless, there are no inertial forces in the measurement signal, even with rapid speed changes, which would be the case with a shaftattached torque sensor. The test shaft as the counterface of seal is fixed via a clamping element and is mounted on a horizontal spindle. The connection to the motor is given by a clutch and optionally a gearbox. The design allows very good accessibility to the seal and test shaft. The friction torque of the seal is measured as well as the controlled rotation speed up to 6000 rpm without gearbox and the temperature at a point on the seal close to the contact using an infrared thermal sensor. To vary the ambient temperature, a heated thermal box can be fitted over the seal and seal support, whereby the Science and Research 28 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0014 very high temperatures resp. very low sliding velocities, at which no suitable lubricating film builds up and higher viscosities resp. lower temperatures lead to lower frictional torques, cp. Teichert [8]. 3 Test rigs for high temperatures There are many test rigs for shaft seals that seal against an oil bath and can therefore be used for applications such as gearbox seals. Here, the temperature at the seal can be influenced very well by influencing the oil temperature, as the heat is primarily dissipated to the oil via the shaft. In the case of minimum oil lubrication or grease lubrication, the heat must be dissipated to the environment via the shaft and the seal by convection. Then, the ambient temperature must be varied using a thermal box, see Figure 3, to analyse the influence of temperature. Frequency [Hz] Complex modulus [MPa] Friction coefficient [-] Velocity [m/ s] Measurement Simulation Figure 2: Analysis of the influence of frequency resp. temperature on the viscoelastic material properties, here the complex modulus, (left) and hysteresis friction coefficient in dependence of velocity (right) Test head Bearing block Drive Test head Bearing Drive Base Counter part Seal k Seal support Thermal isolation Torque sensor Clutch Motor Ther Thermal box M Bearing block Figure 3: Photographs and sectional view of the structure of the RWDR seal test rig with optional thermal box (dashed line) according to Berndt [6] torque sensor remains outside the thermal box and is therefore not heated. The thermal box is indicated by a dashed line in Figure 3 and enables ambient temperatures of up to 120 °C. By measuring the temperature at the torque sensor, a temperature correction of the measured values is implemented. For higher temperatures up to above 250 °C, a new hightemperature seal test rig has been developed that realises the temperature via a thermal chamber with adapted control, see Figure 4. Here, the drive is connected via a toothed belt and a long vertical shaft that extends into the thermal box from below. The application-specific test specimen has an axial running shaft seal and is mounted above the shaft. The torque support of the seal is provided by a long casing around the shaft and is connected to the torque sensor outside the thermal box via a thermal insulation. Holes in the casing and suitable insulation ensure that the shaft and casing expand to approximately the same extent when heated. As a result, the overlap and the axial preload of the axial running seal remain largely unchanged. The design can also be used for radial shaft seals with minor modifications of seal support and conterface. Due to the long shaft with bearings only outside the thermal chamber, operation is possible at low and medium rotation speeds below the critical rotation speed and precise balancing of the shaft is necessary. 4 Experimental investigations Using the thermal box with the first test rig, Figure 3, tests were carried out on radial shaft seals (RWDR) without springs at ambient temperatures of 25 °C, 40 °C and 60 °C. Figure 5 top shows for the 3 ambient temperatures the temperatures at the seal which are mainly shifted in parallel. The friction-specific heating can be added to the ambient temperature for a rough estimate. However, a closer look shows slight differences in the distances between the three coloured curves. Three lubricating oils with very different viscosity classes (ISO VG 32, ISO VG 100, ISO VG 460) were analysed. The corresponding measured friction coefficients, see Figure 5 bottom, show a strong change in the friction characteristics with temperature, especially for the oil (ISO VG 460) with high-viscosity. A first example of the investigation on the high-temperature seal test rig is shown in Figure 6 using an axially running seal. In this test the seal run without lubricant. After running in for 120 minutes at room temperature without heating, the thermal box was heated up from 140 °C to 210 °C in 10 °C steps. For this, a rotation speed of 60 rpm with periodically changing direction of rotation was set. The alternating direction of rotation was used to compensate the temperature-dependent zero-point drift of the torque sensor, assuming equal friction torques in both directions. For this purpose, the mean value of the measured friction coefficient in both directions was subtracted from the friction torque signal. Alternatively, the sensor can be calibrated as a function of the temperature in the thermal chamber. It was found that the friction torque decreased significantly when heated to 140 °C for the axially running shaft seal used. Further heating led to an additional decrease in friction torque. At the thermal chamber temperatures of 200 °C and 210 °C, the frictional torque also Science and Research 29 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0014 Seal Casing Shaft Thermal chamber Torque sensor Drive system Thermal isolation Figure 4: Photograph and sectional view of the structure of the high-temperature seal test rig 5 Conclusion Elevated temperatures are particularly critical in the case of minimum oil lubrication or grease lubrication. The influence of both moderately increased ambient temperatures on the sealing contact temperature and the seal Science and Research 30 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0014 initially decreases during the 60-minute measurement at a constant temperature, but then increased again. This is an indication of changes in the tribological system and may be caused by changes in the seal material at the high temperatures. These can be investigated in more detail by analyses after the test run. Velocity [m/ s] Friction coefficient [-] Temperature [°C] Velocity [m/ s] Velocity [m/ s] Velocity [m/ s] Velocity [m/ s] Velocity [m/ s] Friction coefficient [-] Temperature [°C] Friction coefficient [-] Temperature [°C] Figure 5: Temperature (top) at the seal contact and coefficient of friction (bottom) of a radial shaft seal without spring at three ambient temperatures (blue: 25 °C, green: 40 °C, red: 60 °C) for three different oil viscosities (left: ISO VG 32, centre: ISO VG 100, right: ISO VG 460) according to Berndt [6] Figure 6: Frictional torque (only maxima and minima of torque are shown) and controlled temperature curve of a temperature rise test with cyclic changing rotation directions on the high-temperature seal test rig of an axially running shaft seal, here tested without lubricant friction as well as greatly increased temperatures on the frictional torque could be investigated using two specially test rigs. Several influences on the tribological effects of the temperature were shown. The friction torque curves showed initial changes of the seal properties at high temperatures above 200 °C. Literature [1] Freudenberg Sealing Technologies GmbH: Technisches Handbuch, Kapitel 1: Simmerringe und Rotationsdichtungen, 2015. [2] Heinz K. Müller: Abdichtung bewegter Maschinenteile: Funktion - Gestaltung - Berechnung - Anwendung. Waiblingen: Medienverlag Müller, 1990. [3] Tim Leichner: Prognose der Dichtlippenfolgefähigkeit von RWDR bei dynamisch verlagerter Welle. Maschinenelemente- und Getriebetechnik-Berichte 9. Technische Universität Kaiserslautern, 2012. [4] Tobias Engelke: Einfluss der Elastomer-Schmierstoff- Kombination auf das Betriebsverhalten von Radialwellendichtringen. Dissertation. Leibniz Universität Hannover, 2011. [5] Kathrin Ottink: Betriebsverhalten von Wälzlagerschutzdichtungen: Experimentelle Untersuchungen und Berechnungsansätze. Dissertation. Leibniz Universität Hannover, 2014. [6] Christian Berndt: Tribologie von Radial-Wellendichtungen. Dissertation. Technische Universität Bergakademie Freiberg, 2019. [7] Ringo Nepp: Experimentelle und theoretische Adhäsionsanalysen an Reifen-Profilklötzen und O-Ring-Dichtungen. Dissertation. Technische Universität Bergakademie Freiberg, 2017. [8] Robert Teichert: Tribologie von fettgeschmierten Radialwellendichtungen. Dissertation. Technische Universität Bergakademie Freiberg, 2022. [9] Malcolm L. Williams, Robert F. Landel, John D. Ferry. The temperature dependence of relaxation mechanisms in amorphous polymers and other glass-forming liquids. Journal of the American Chemical Society 77.14 (1955), pp. 3701-3707. Science and Research 31 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0014 suitability of a lubricant for their application. They ask the bearing and / or gearbox manufacturers to release a lubricant. To release a lubricant, an approval is required and mostly it is related to minimum requirements for the lubricant candidate. The choice for a suitable lubricant includes the requirement for the properties of minimizing friction, remove heat and provide sufficient viscosity, to separate contact surfaces, without harming the gears and bearings. Based on these requirements, test criteria and requirements are defined. For rolling bearings and gears, an example for those requirements can be found in [18]. What is missing are requirements for the damage mode WEC. The proposed FE 8 test rig A lot of research results are available for the axial cylindrical roller bearing 81212. In this paper 139 data files are available for the 81212 and 2 for the 81206, reported in 7 publications. See table A-1. With WEC, 101 bearings failed and without WEC, 40 bearings are reported. For the assessment of the possibility of occurrence of WEC, the calculation model [3, 4, 5] is used. With this model it is possible, to calculate a mechanical basic safety for the tested bearings, based on design, load, bearing capacity and speed, which can be used, to compare test results among themselves, a so-called compa- Science and Research 32 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0015 Introduction Since nearly 2011, WEC bearing damages have been observed. WEC damages are reported for instance by Johan Luyckx [1] in the industrial application wind turbine gear unit. In his presentation, failure rates of up to 70 % during the first 2 years are mentioned for non-coated bearings. These damages are identified as “White etching Cracks“ so called WEC. First research investigations with respect to WEC failures are published in 2014 by the Forschungsvereinigung Antriebstechnik FVA [2]. In the Bearing World Journal Vol. 5 (2020) [3], a knowledge based analytic and AI suitable calculation method is published by Leimann, to evaluate the risk of WEC occurrence in industrial applications. The method can adjust the quality of the results in a self-learning process by each added test result. WEC damages on rolling bearings, their damage patterns, hypotheses of origin and influences on the origin are deeply discussed in the GfT presentation, reported in 2023 [17]. In this presentation, also design proposals are done to reduce the risk of the occurrence of WEC, for instance, the choice of bearing material and coatings as black oxide. With respect to the influence of lubricants, a lot of research tests are carried out. Unfortunately, a direct assignment for these lubricants between technical or chemical component data as the type and quantity of chemical elements is, with view to the state of the art of research, hardly possible. There is a lot of research done to find a direct assignment of chemical elements and their volume share to the occurrence of WEC, but until now, there is no clear result. Also, field observations of damaged bearings cannot help to differentiate lubricants with respect to their behavior of creating or not creating WEC, because the operating and ambient conditions differ a lot during the runtime and influences as water or current or voltage go through are difficult to capture in the field. End customers and OEMs expect from bearing, gearbox, and lubricant manufacturers recommendations about the Thoughts on a standardized FE-8 Test for the assessment of the WEC-carrying capacity of lubricants in rolling bearings Dirk-Olaf Leimann * Presented at the GfT Conference 2024 For an amount of 141 WEC-Tests on axial cylindrical roller bearings on a so-called FE 8 test equipment, a mechanical WEC safety is calculated from the time to failure value compared with the L 10h life. From these values, a factor is derived to get an idea about the influence of the tested lubricants on WEC forming. Furthermore, a study on relevant influence parameters as load and speed is done. Out of these data, a proposal is done for a standardized WEC-test for lubricants. Keywords WEC-Tests, axial cylindrical roller bearings, wind turbine gearboxes, WEC damage, FE 8 test rig Abstract * Dipl.-Ing. Dirk-Olaf Leimann, Düsseldorfer Straße 4, 47441 Moers rative reliability or safety S WEC . Out of this comparison it is possible, to define test conditions, which are suitable for lab testing at research institutes and test labs from lubricant suppliers on available FE 8 test equipment. The paper presents results of the mechanical WEC safety S WEC for the 141-bearing data in the table A-1. It is visible that, without taking the lubricant into account, all bearings would not fail. With the time to failure of the WEC damaged bearings, a second factor is calculated for the bearings with WEC damage. Out of these results a factor is formulated, to capture the influence of a lubricant. Furthermore, a parameter study is done with variations on load, speed, and number of rollers to determine suitable test parameter. To complete the investigations on influence factors, the 141 data are evaluated with respect to lubricant technical and chemical properties. Taking all results into account, a standardized FE-8 test for the assessment of the WEC-carrying capacity of lubricants in rolling bearings is proposed to discuss. The proposed test parameter reflects to the axial load, speed, number of rollers, oil flow, material, oil viscosity at 100 °C and test temperature. Additional test parameter from research results should be discussed to add. The so-called FE 8 test design (Image 1) is a well-known and accepted test equipment with a high availability at research institutes and industry to test lubricants regarding their suitability for the use with rolling bearings. On FE 8 test rigs, test bearings can be axial cylindrical roller bearings 81212, taper roller bearings 31312 and angular contact ball bearings 7312. Information about the FE 8 test rig with more details can be found in the GfT data sheet “GfT FE8 N 060” [6]. The mentioned bearing types represent a large spectrum of industrial applications. The standard material of these bearings is 100 Cr 6. In the research investigation 707 I of FVA [2] a lubricant “Low Ref SAE30” is used with a charge 1 and 2 and an axial cylinder roller bearing 81212 on a FE 8 test rig. By using this lubricant, WEC damages could be reproducible produced with the axial cylindrical roller bearing. Tests with a radial cylindrical roller bearing with this lubricant and similar load, but higher speed conditions, could not produce WEC damages. Especially with research on wear behavior, WEC damages could not be observed on FE 8 test rigs, even with a big range of different lubricants all kind of. For instance, in the cited FVA investigation Nr. 327 „Entwicklung experimenteller Grundlagen für eine schmierstoffabhängige Verschleißlebensdauerberechnung für Wälzlager (Ölkennwerte für Wälzlager)“ [7], 10 very different lubricants were examined with bearing types 81212, 31312 Science and Research 33 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0015 Image 1: FE 8 test rig with angular contact ball and axial cylindrical bearing [2,6] Table 2 shows, that non coated bearings have failure rates up to 70 % in between 2 years run time. Problem description and existing recommendations for rolling bearing tests The GfT presentation “White Etching Cracks” (WEC) [17] lists possible influences to generate WEC damages. Mainly, these could be current or voltage go through, hydrogen embrittlement and chemical elements or compounds. In tests, a correlation between damage occurrence and quantity of hydrogen, amperage or current voltage could be found. Unfortunately, direct correlations for lubricants between chemical elements or compounds and quantities could rarely be seen, because of too much possible components and elements and their chemical possibilities. So, the only judgment could be, WEC damage occurred or not. From the IEC 61400-4 [8] it is known, that a lubricant is suitable for rolling bearings in the wind turbine application, if the lubricant fulfills the recommendations, shown in the tables E5 and E6. Rolling bearings should fulfill those requirements with respect to wear and fatigue behavior, determined on FE 8 tests. Table 3 is a summary of test methods, load parameters and requirements. The load parameters are like those, reported in table A-1 in the annex. Similar recommendations from the industry and specifications are shown in table 4. Based on the approach of the IEC 61400-4, the availability of test specimen and test equipment, a sufficient data base on WEC tests for the axial cylindrical roller bearing 81212 with and without WEC damages and the possibility, to cover the mechanical basic safety, it is obvious, to define test parameters for a WEC lubricant test on FE 8 test rigs. Science and Research 34 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0015 and 7312 without reported subsurface damages. Contact stress and test duration are comparable to the data from [2]. The oil viscosity at 100° C was between 7 and 16 mm 2 / s. Speed values were lower with a factor of 4 to 40. There were also test times with a duration of more than 500 hrs. In literature it can also be found, that WEC damages can also be generated on test rigs with radial bearings. But these test rigs differ a lot from each other in design, bearing type and size and are not standardized as the FE 8 test rig [21] with the bearing 81212. For wind turbine gear boxes, a list of recommendations for the assessment of gear lubricants are mentioned in tables E5 and E6 [8] with respect to FE 8 tests and the use of the lubricant in the wind turbine application. The calculation method to evaluate the risk of WEC occurrence In the cited papers [3, 4, 5], there are detailed information’s about the applied calculation method. The method is a knowledge based analytical model, based on real field data and results from research test data. The method compares permissible contact stresses with occurring contact stresses and from those, WEC safety factor S WEC is formed for the assessment of the probability of WEC failures. This method is used to find design and mechanical test parameter as load, speed, contact stress, nominal life time and number of rollers, which enables to find conditions, that could make a serious influence of a lubricant visible. The safety factor judges the risk of the occurrence of a WEC damage. WEC damages on Wind Turbine Gear Boxes Johan Luyckx [1] reports in 2011 serious bearing damages due to WEC on wind turbine gearboxes. He compares black oxide coated bearings with non-coated bearings. ! " # $# # $# % & Table 1: Definition of the safety factor S WEC with respect to the probability of occurrence Table 2: Failure rates versus gearbox population Image 2: WEC damages  The data pool Out of various publications, available data for at least 141 WEC tests were collected. These files include design data, load data, data about the lubricant and other informative data. In total 49 specific data for each tested bearing were available. The main data regarding design, load, speed, and lubricant are summarized in the table annex A-1. With the data from the 141 bearing tests, the following tables and diagrams are extracted. Compared are the results of all 141 data files with 101 data files from bearings with WEC and the 40 files without WEC. The life time factor S time to failure in table 5, 6 and diagram 2 is calculated from the time to failure according to the equation (1). (1) with: L t time to failure in hrs L 10 h calculated life in hrs The diagrams show the results of the cases with WEC. Table 5 shows the minimum, maximum and average values for the main parameters and an estimation of the influence of the lubricant. The 141 data files include 139 data sets for the axial cylindrical roller bearing 81212 and 2 sets for the bearing 81206. 137 tests were done with an oil flow between 0,1 and 0,25 l / min and 4 tests were done with an oil flow of 1,2 l / min. With the higher oil flow of 1,2 l / min, WEC damages were not observed. 1 3346 73879 4: ; 7 7 7 7 Science and Research 35 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0015 Table 3: Test recommendations in IEC 61400-4 “Design requirements for wind turbine gear boxes” Table 4: Examples for the recommendations from industry for tests with FE 8 [19,20,21] 346 73879 4: ; 7 Table 5: Evaluation of the 141 data files mentioned in [8] and failure rates mentioned in [1]. With these estimated data, the average damage factor is 1,90 (see diagram 2). That means, that a FE 8 test could representative and be used, to give a judgment on the suitability of a lubricant with respect to WEC damages. Science and Research 36 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0015 The results of the 101 data sets contain 22 different lubricants, where WEC damages could be created. As visible in the diagram 2, the average of the estimated damage factor is about 2,1. In table 6, an estimation of the damage factor is done for a three-stage planetary gear unit for the application wind under the conditions Table 6: Possible WEC damage factor with the assumptions from [8] and [1] Diagram 2: Results of the possible influence of lubricants to create WEC Diagram 1: Data and results for the 101 bearing tests with WEC damage Design and load influence The evaluation of table 5 shows, that WEC damages can occur at contact stresses between 1750 and 2219 MPa and loads between 25 and 100 kN. In table 7 in IEC 61400-4 [8] it is recommended, that the maximum contact stress should not exceed values between 1300 and 1650 MPa. Looking to the speeds in table 5, it can be recognized, that they are in the range of table 6 and that the average and minimum values of non WEC test sets are higher, than with WEC. In the research project FVA 707 I [2] also tests with the radial cylindrical bearing NU 206E in combination with the lubricant “Low Ref SAE30-2” were carried out. While using the same contact stress and same lubricant, no WEC damage could be created. The axial cylindrical roller bearing 81206 had 12 rollers and the bearing 81212 was tested with 15 rollers. In other cited documents, also 19 rollers are reported for the 81212. Influence of a lubricant The cited publications on WEC tests on FE 8 test rigs show, that the lubricant could have a serious influence for WEC creation. How these lubricants behave in the field is difficult to estimate because of variable ambient and load conditions. Data for field experience are rarely published. One very good description on field experience is done by O.L. Jensen [15], but more detailed information’s are not available. For 76 from 141 WEC tests, data about the lubricant and the possible influence of mechanical data and data about chemical elements and compounds are available. 65 bearing tests failed with WEC and 11 tests had no WEC damages. Table 7 shows a summary of some of those more specific lubricant data. Suitable test parameter Out of an analysis of all data it turns out, that a test bearing 81212 with 15 rollers is suitable for a standardized test on a FE 8 test rig with load parameters, like research test and field experience. The proposal is supported by the calculated WEC safety S WEC . The variation of parameters and their results are shown in table 8. From table 8 it can be seen, that a test speed of 350 rpm in combination with a bearing load of 60 kN, resulting in a contact stress of 1750 MPa, are suitable test parameters. • Bearing 81212 • Number of rollers 15 • Oil inlet volume 0,2 l/ min • Material 100 Cr 6 Science and Research 37 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0015 Table 7: results of the evaluation of specific lubricant data Table 8: Suitable test parameters for a standardized WEC test [7] N. van de Sandt et al, „Entwicklung experimenteller Grundlagen für eine schmierstoffabhängige Verschleißlebensdauerberechnung für Wälzlager“, Vorhaben FVA 327, Heft 665, 2011 [8] NN, Wind Turbines - Part 4, Design requirements for wind turbine gear boxes, IEC 61400-4, Genève, 2012 [9] H.K. Danielsen et al, “Accelerated White Etch Cracking (WEC) FE8 type tests of different bearing steels using ceramic rollers”, Wear 494-495 (2022) 204230 [10] H.K. Danielsen et al, “FE8 type laboratory testing of white etching crack (WEC) bearing failure mode in 100Cr6”, Wear (2019) 202962 [11] J.Loos, „Einfluss der Reibbeanspruchung auf die WEC- Bildung in Wälzlagern“, Tagungsband GfT Fachtagung, Göttingen, 2014 [12] A.D. Richardson et al, “The Evolution of White Etching Cracks (WECs) in Rolling Contact Fatigue-Tested 100Cr6 steel”, Springer Verlag, Tribology Letters (2018) [13] H. Surborg, „Einfluss von Grundölen und Additiven auf die Bildung von WEC in Wälzlagern“, Dissertation Universität Magdeburg, Shaker Verlag 2014 [14] F.G. Guzmán, White Etching Cracks (WEC) in ölgeschmierten Wälzkontakten, Verlag Mainz, 2020, ISBN 978-3-95886-340-8 [15] A. Ruellan, K. Stadler, “Can oil chemistry accelerate WEC-associated damage in thrust & radial bearings? ”, Bearing World 3 rd international FVA-conference, 2020 [16] O.L. Jensen et all, “Prevention of “White Etching Cracks” in rolling bearings in Vestas wind turbines” Tagungsband CWD ATK, Aachen, 2021 [17] F.G. Guzmán et all, “GfT-Positionspapier „White Etching Cracks - WEC“ Tagungsvortrag 63. Tribologie-Fachtagung, GfT, Göttingen 2022 [18] D. Leimann, “Hansen selection criteria for lubrication oils for gearboxes in wind turbines” Tagungsband 17 th International Colloquium Tribology, TAE, Esslingen, 2009 [19] NN.: Druckschrift 7300de, “Getriebeschmierung und Getriebekonservierung”, Flender Bocholt, 2022 [20] NN.: Druckschrift PI 4-1104, “Produktinformation Fuchs Renlolin Unisyn CLP“ Mannheim [21] NN, DIN 51819-3 „Prüfung von Schmierstoffen - Mechanisch-dynamische Prüfung auf dem Wälzlagerschmierstoff-Prüfgerät FE8 - Teil 3: Verfahren für Schmieröl - einzusetzende Prüflager: Axialzylinderrollenlager”, DIN-MEDIA 2016 Science and Research 38 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0015 Proposal To evaluate a lubricant regarding the suitability in an industrial application and to give a judgment about the WEC behavior, the test parameters from table 8 should be further discussed. The proposal is based on the best practice approach of the IEC 61400-4 [8] for recommendations on lubricants for the use in wind turbine gearboxes Citation from [8]: “Wind turbine gearboxes can be quite varied in their design and configuration. As such, lubrication requirements may vary among gearboxes and also by their operating environment, i.e., ambient conditions, duty cycle, etc. The information contained in this annex is designed to provide users (end users, owners, turbine manufacturers, and component suppliers) with a guideline for a minimum level of lubricant performance for this application. Additionally, guidelines for condemning limits of selected lubricant parameters are offered based on experience in the industry.” So, a judgment could be: If a test according the parameters from table 8 is carried out on a FE 8 test rig and the lubricant did not lead to the formation of WEC damages, the minimum conditions of the test are fulfilled and the tested lubricant is suitable for the application. Literature [1] J. Luyckx, “WEC failuremode on rollerbearings”, VDI- Wissensforum, Tagung Gleit- und Wälzlager Schweinfurt, Düsseldorf 2011, ISBN 978-3-18-092147-1 [2] C. Bongardt, M-O Özel et al The Evolution of White Etching Cracks (WECs) in Rolling Contact Fatigue-Tested 100Cr6 steel, Risse auf Lageringen, Gefügeveränderungen in Wälzlagerringen mit Rissen als Folgeschaden, Vorhaben FVA 707 I, Heft 1121, 2014 [3] D. Leimann, “Calculation Method to Evaluate the Risk of WEC Occurrence in Industrial Applications”, Bearing World Journal Vol. 5 (2020), VDMA, Frankfurt 2020 [4] UK Patent Applikation GB2585272A “Comparable stress for rolling bearings”, Intellectual Property Office, Newport, United Kingdom, 2021 [5] Offenlegungsschrift DE102021128876A1 “Vergleichspannung für Wälzlager 2”, Deutsches Patent und Markenamt, München, 25. Mai 2022 [6] N 060 mod 1, GfT Arbeitsgruppe „Datenbank Tribologische Prüfstände“ - Datenblatt FE 8 Prüfgerät, Science and Research 39 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0015 Annex Table A-1: bearing test data for 81212 and 81206 Science and Research 40 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 Introduction: brake fluid and challenges noise and wear Passenger cars and light commercial vehicles are equipped with hydraulic brake systems. A hydraulic medium transmits and converts the braking force from the brake pedal in the brake master cylinder into hydraulic pressure, which is routed via brake lines to the wheel brakes and there converted back into braking force. Vehicle-dynamics control systems such as the antilock braking system ALC or the electronic stability program ESP can additionally modulate the hydraulic pressures at each wheel to prevent the wheels from locking through pressure reductions or to stabilize the handling through additional pressure build-ups. Brake fluids must comply with stringent requirements to ensure reliable brake-system operation. Their properties are defined in various standard requirements which are very similar in terms of content (SAE J1703 [1], SAE J1704 [2], FMVSS 116 [3], ISO 4925 [4], JIS K2233 [5]). Exclusively for the vehicle operating medium brake fluid, the performance requirements of FMVSS 116 (Federal Motor Vehicle Safety Standard) became mandatory in the USA and often serve as a universal reference. In this standard, the US Department of Transportation (DOT) - comparable with other standards - has defined specific ratings for salient properties [7]. A list of criteria enables material compatibility (e.g. corrosion limits for metals, swelling limits of rubber parts (in case of polyglycol based brake fluids, EPDM is the preferred rubber material),…) and provide force/ pressure transfer function over a large application range (viscosity & boiling point requirements defining the brake fluid classes,…). Improvements in brake fluid standardization to avoid noise & wear problems Michael Hilden, Gerd Dornhöfer, Harald Dietl* Presented at the GfT Conference 2024 Brake fluids offer large temperature ranges of incompressible fluidity for dynamic braking performance. Modern brake fluids combine increased boiling points in “dry” (new fluid) and “wet” (used fluid after few years in vehicle) conditions with reduced cold viscosity to achieve minimized vapor lock risks with more dynamic pressure build up performances. However, reduced lubrication performance risks have been observed within the development of first modern brake fluids. Since sealing rings in braking systems are “lubricated” by brake fluid and form tribo-systems, lubrication performance of the brake fluid is essential in vehicle application and is currently integrated into brake fluid standardization. Automated driving and increasing noise requests due to electrification will further increase tribological requests for the pressure medium. In order to avoid future wear and noise problems, it is essential to develop and validate appropriate test methods to assess and prescribe lubrication performance of brake fluids. Within the project TriNoWe - funded by the Federal Ministry for Economic Affairs and Energy (BMWi) - the impact parameters, stress conditions and their dependencies based on a deep tribo-system analysis were identified to develop appropriate test methods. A valid noise lab test method - provoking the product behavior in a lab test - was developed as DIN 51834-5 and was already introduced as pass/ fail into international brake fluid standards. For wear, an appropriate test method based on identical test specimen and comparable tribometer use (but with incline setting, high temperature and longer load run time with load increasing starts and stops) - applicable after performed noise test - was developed as DIN 51834-6. It is in release phase by finished round robin tests to be also introduced into brake fluid standardization soon. Keywords Brake fluid, test method, lubrication, noise, stick-slip, friction, SRV, wear, surface conditioning Abstract * Dr.rer.nat. Michael Hilden Robert Bosch GmbH Robert Bosch Allee 1, D-74232 Abstatt Dr.rer.nat. Gerd Dornhöfer Robert Bosch GmbH Robert-Bosch-Campus 1, D-71272 Renningen Dr.rer.nat. Harald A. Dietl BASF SE Carl-Bosch-Strasse 38, D-67056 Ludwigshafen am Rhein The applications do also contain rubber (EPDM) sealing rings for moved metal pistons, e.g. in the master cylinder of the booster or in the pump elements of the electronic stability program unit modulating the brake pressure. However, no criteria for lubrication are contained in the brake fluid standards (beside a stroking test in FMVSS 116 not applied in other standards due to poor availability of old test specimen). Challenge noise Few brake fluids are known to provoke noise when applying the brake pedal by generating stick-slip oscillations. Such noise challenges typically disappear, if the brake fluid is changed to another type. Thus, brake fluid suppliers try to modify such brake fluids in order to avoid the noise challenges typically by performing noise tests in selected products already within fluid formulation design. For electrical vehicles without masking combustion engine, the noise level in the vehicle is reduced and the driver will have more awareness and notice noise challenges more disturbing. Thus, appropriate brake fluids have to avoid noise issues in field (see upper line in Figure 1). Challenge wear Modern brake systems shift the brake force application from the master cylinder (partly-) automatically to the electronic stability program, e.g. if sensors detect the need for vehicle deceleration. In brake fluid performance tests of a new modern brake fluid with an increased load scenario, wear of sealing rings of the pump elements was detected by different brake component suppliers in product tests. In order to avoid such wear issues in field, a certain minimum lubricity quality is required for the brake fluids - but currently not contained in the criteria list of the standards (see lower line in Figure 1). Targets of SAE & ISO TF brake fluid lubrication and public funded project TriNoWe The trends of electrification and automated driving underline the need for a minimum lubricity requirement of brake fluids. In order to assess this lubricity product-independent, appropriate lab tests have to be developed and validated in order to introduce them into the brake fluid specifications. Two active international committees for brake fluid standardization, SAE and ISO, did agree to integrate such criteria and did from the joint SAE & ISO task force brake fluid lubrication defining exactly these tasks in 2017. Many partners of the SAE & ISO task force did initiate the public funded project TriNoWe to focus on the tasks. Figure 2 presents the TriNoWe tasks on high flight level: For the two challenges noise (upper line) and wear (lower line), the product behavior (left block) has to be rebuild by simplification and abstractions (middle block) by basic lab tests moving rubber (in brake fluid applications EPDM (Ethylen-Propylen-Dien-rubber) against metal bodies (right block). Finally, product challenges have to be provoked in simplified lab tests. Therefore, the project TriNoWe -publically funded by BMWiidentified the impact parameters, stress conditions and their relationships based on a deep tribo-system analysis. Appropriate test bodies have been identified or developed for the lab tests, see right hand side of Figure 2. Science and Research 41 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 Figure 1: Challenges noise and wear with provoking trends to define fluid lubrication targets Science and Research 42 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 Noise lab test DIN 51834-5 Figure 3 shows basics of the developed noise lab test DIN 51834-5. In the upper line center, the test body scheme is shown as metal ball and rubber (EPDM) disc contact in the tribometer setting. On the right hand side, a noisy fluid test is shown and on the left hand side an ok and not noisy fluid test of the reference fluid RF ISO defined in ISO 4926 [6], identically applied as RM 66-07 by SAE as reference fluid (RF ISO = RM 66-07). Only difference in the test and pictures is the test fluid itself (no picture difference visible), but audible noises occur that are also measurable and assessable by stick-slip oscillations of the friction coefficient on the right hand side. Thus, within the developed setting, the system noise behavior is provoked in the tribometer test, measurable by stick slip oscillations of the coefficient of friction µ(t). Consistent to the audible noise with fluid RF 31, the stick-slip effect is also noticeable within the measured friction force signal (or friction value µ(t)) as severe oscillations in case of stick-slip as shown in Figure 4. In the left diagrams, where two cycles of the tribometer test for RF 31 (red, upper line) are compared to RF ISO (green, lower line). Oscillations occuring in µ(t) are mathematically evaluated by subtracting the filtered data obtaining Δ as remaining oscillations and determine their deviation as an appropriate measure σ(Δ). Figure 5 shows the high repeatability of the test results in a statistical study performed by Bosch: Three operators did test 10 test fluids (5 fluids RF ISO and 5 fluids RF 31 = RF G in double repetition, i.e. 10 ok sample tests and 10 not ok sample tests per operator). The high reproducibility is finally possible by selection of high quality metal ball surfaces (G5), valid EPDM sample Figure 2: Tasks of TriNoWe: develop valid lab tests for noise and wear reflecting product results Figure 3: Differentiation between not noisy and noisy fluid in the tribometer with identical set up y p μ( ) Science and Research 43 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 Figure 4: Measured differentiation between RF 31 (upper line) and RF ISO (lower line) and mathematical evaluation method σ(Δ) (left measurement, middle method, right complete test run results in 15 different working points) Figure 5: Bosch performed a statistical study of noise test DIN 51834-5 Figure 6: Results of R2TN: Round Robin Test Noise (averages of 7 international labs with sigma on y-axis as valid pass/ fail criterion) within a component endurance test at 100 °C, applied after observed wear with a selected brake fluid. The wear pattern of the sealing ring is measured by the change of the inner sealing diameter referenced in an optical diagram between original inner ring diameter and outer metal piston diameter. This test method was applied as component test at Bosch to assess the lubrication behavior of the brake fluids with respect to wear, but it takes a few weeks run time and high efforts and is product based. Figure 8 explains an essential difference between product and tribometer tests observed during TriNoWe: in the product, the sealing ring separates the fluid-side from the air-side and the wear is typically starting from the air-side. Thus, the root cause for the wear is combined with the missing availability of lubricating brake fluid in the contact. Science and Research 44 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 (DIK recipe according to Appendix A of DIN 51834-5) and its surface roughening as PDR, where a sharp metal tool harmonizes the EPDM surface of the test discs in dry condition before the fluid test. A round robin test noise (R2TN) was performed within the SAE&ISO TF brake fluid lubrication in collaboration with DIN 51834. The valid differentiation between noisy fluids (red points with high sigma and increased static friction coefficient (SFC, standard tribometer output) and not noisy fluids (green points with low sigma values) below the pass/ fail criterion 0,005 was confirmed by seven international lab results (see Figure 6). Wear lab test DIN 51834-6 Figure 7 explains the Bosch procedure to assess the wear behavior of brake fluids by applying a defined load Figure 8: Difference between tribometer and product wear tests Figure 7: Component test procedure for wear at Bosch, applied as assessment for brake fluids However, in the tribometer “point contact” without real sealing behavior, the availability of brake fluid is always given and never limited. Thus, the dominant wear mechanism is not rebuilt sufficiently within this test system. Modifications of the tribometer test towards component tests behavior were required. Bosch did propose to rebuild the separation between fluidand air-side in the tribometer by applying the test in incline settings of the tribometer with many starts and stops. Severe wear pattern could be provoked in such test settings already after short test times. Furthermore, the beginning of “lack of lubrication” can be measured in the tribometer by a jump in the friction force. Thus, the applied load cycles until this jump happens could be applied as simplified assessment criteria for wear. Figure 9 shows the valid and required enrichments of the tribometer test and the valid fluid differentiation. Essentially for the success was the development of an enhanced test bath as required adapter and a precise introduction of only 0.4 mL test fluid amount (yielding half EPDM disc covered in new test fluid bath). It also contains the important detail, that also further suppliers tested the wear reference fluid MTG with strong wear in their component tests. Figure 10 compares the measured friction data and wear profile of the EPDM discs. In the upper graph, the measured friction values jump between 0 (result delivered during breaks) and measured friction value resulting in a black area. The next graph below shows the 200 s averaged friction values as profile line, with its derivate in the graph below to detect a friction jump when exceed- Science and Research 45 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 Figure 10: Valid differentiation between wear ok reference fluid (left hand side) and wear reference MTG (right hand side) with friction jump detected already after 2.232 load cycles and significant wear depth on the EPDM test specimen Figure 9: Enrichment of tribometer test to rebuild wear behavior: incline position and start / stop integration on between wear and no wear fluids of 7 international labs with the results of the statistical study in the background in bright colors. Finally, system challenges noise and wear can be provoked by appropriate lab tests with transferable test fluid results as summarized in Figure 13. Figure 14 compares the test criteria and the repeatability of the noise (upper) and wear test (lower line) and the already achieved status for standardization on the left hand side. Since identical test specimen are applied for noise and wear test (both with identical EPDM preconditioning by “PDR”), it is valid to first perform the noise test and Science and Research 46 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 ing a slope limit after a certain number of load cycles to be evaluated during the test run. The wear profile has to be measured as cross section line across the EPDM disc over the center of the PDR window enabling a valid and reproducible measurement of the wear depth d wear as assessment size for pass/ fail criterion. Figure 11 shows the high reproducibility of the results by a statistical study performed by Bosch. Also for the wear test a high reproducibility was achieved: all 30/ 30 tests of ok reference RF ISO do not show wear without friction jump and all 30/ 30 of wear reference MTG do show wear with a friction jump and Figure 12 summarized the round robin test results with valid differentiati- Figure 12: Results of R2TW-2 Round Robin Test Wear (closed 08/ 2024) Figure 11: Analysis of statistical study of wear test DIN 51834-6 afterwards the wear test (cleaning and valid refilling of fluid amount preserved). Test specimen as DIK reference rubber and standard 10 mm diameter metal balls with G5 surface quality are defined in the standards DIN 51834-5 and -6 and are available long term without concerns and challenges of the test specimen of stroking test in FMVSS 116. Finally, the targets of the pfp TriNoWe and of SAE & ISO are reached and fulfilled. Conclusions The project TriNoWE developed a tribometric test procedure in order to assess the potential of brake fluids to generate friction-induced noises in EPDM-metal contacts by determining the potential fluctuations in the frictional response in a wide range of operating conditions. An appropriate EPDM surface conditioning approach is essential to map lab test and product results and to obtain a valid and robust test procedure disclosed as DIN 51834-5. The noise test is finished and already introduced into SAE & ISO brake fluid standards. It succeeds in provoking stick slip noise of product noisy fluids in a defined lab test procedure. The wear test DIN 51834-6 is in closure after performed statistical study and round robin test wear R2TW. Since Science and Research 47 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 Figure 14: Comparison of noise and wear test reproducibility and status of integration into brake fluid standards of SAE & ISO Figure 13: Success of develop valid lab tests for noise and wear DIN 51834-5 & -6 reflecting product results [6] ISO 4926: Road vehicles - Hydraulic braking systems - glycol-based reference fluid. [7] Hilden, M., Dietl, H.: Brake fluids, page 322-325 in: Kraftfahrtechnisches Taschenbuch, 2019, Springer Verlag, ISBN 978-3-658-23583-3 [8] Hilden, M., Robert Bosch GmbH: Future mobility and driving assistance: How to deal with new challenges for brake fluids? Evolution in Motion, BASF’s Fuel and Lubes Conference 2018, October 16-17, 2018. [9] Hilden, M.: Results & follow up of the pfp TriNoWe - methods to assess lubrication performance of brake fluids, Maschinenbau-Kolloquium Hochschule Mannheim, 17.05.2021 (online) [10] Hilden, M.: Public funded project TriNoWe: Development of NOise & WEar standard tests for brake fluids, Vortrag bei der Jahrestagung der Gesellschaft für Tribologie e.V., 09/ 2021, pp.49/ 1-49/ 13, ISBN978-3-9817451-6-0 [11] Hilden, M. and Dietl, H.: Improvements in brake fluid standardization to avoid noise & wear, Vortrag bei ATZ Chassis.tech plus 2022, 06.07.2022, München. Veröffentlichung im Tagungsband, ISBN 978-3-662-70347-2 Science and Research 48 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 DOI 10.24053/ TuS-2024-0016 wear is a much more severe failure mode compared to noise, it should be expected that a comparable reproducible test method for wear will also be introduced into SAE & ISO brake fluid standards as requested already in 2017 by charging the SAE & ISO TF brake fluid lubrication to develop the two appropriate lab test method for standard introduction. References [1] SAE J1703: Motor Vehicle Brake Fluid. [2] SAE J1704: Motor Vehicle Brake Fluid Based Upon Glycols, Glycol Ethers and the Corresponding Borates. [3] FMVSS 116: Federal Motor Vehicle Standard No. 116: Motor Vehicle Brake Fluids. [4] ISO 4925: Road vehicles - Specification of non-petroleum-base brake fluids for hydraulic systems. [5] JIS K2233: Non-petroleum base motor vehicle brake fluids. News 49 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 In July 2024, the Young Tribologists Working Group of the German Society for Tribology (GfT) organized the 7 th Young Tribological Researcher Symposium at the Chair of Engineering Design at Friedrich-Alexander- Universität (FAU) Erlangen-Nürnberg under the patronage of Prof. Dr.-Ing. Sandro Wartzack. The symposium provided a platform for young researchers to showcase their new ideas and ongoing research projects. A total of 11 presentations were delivered in English, covering a wide range of topics, from elastohydrodynamic simulations to the investigation of implants. The latter was connected to the plenary talk by Kevin Neusser (KTmfk) entitled “Nowear to be found”. Current results regarding the coating of endoprostheses were presented in an entertaining and scientific manner. Following the presentations, during coffee breaks and lunch, the approximately 25 participants had the opportunity to engage in discussions to delve deeper into the various topics. The award for the best presentation, as voted by the audience, went to Eleonor Carberry from RWTH Aachen University for her contribution titled “Solid lubrication with PTFE in rolling contacts” The symposium was accompanied by a social dinner and an excursion to PETER BREHM GmbH - a manufacturer of medical implants and prostheses. The Young Tribologists would like to express their heartfelt gratitude to all participants for their contributions and, especially, to the sponsors Optimol Instruments Prüftechnik GmbH, Oerlikon Balzers AG and Schaeffler Technologies AG & Co. KG for their support of the YTRS. Reserve the date: The next YTRS will take place on July 21-22, 2025, at the Chair of Machine Elements, Gears and Tribology of the University of Kaiserslautern-Landau (RPTU). Stay tuned! Gesellschaft für Tribologie 7 th Young Tribological Researcher Symposium (YTRS) News 50 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 Österreichische Tribologische Gesellschaft ! ""'(./ #) ) .0"#%&.1)#%&*+,+-2)"+'&$#2 ! ! ! ! ! ! "#$%&'()%*+'$,-,.$ ! "#$%&%'()#*)+*,-./ "()0*,)12.20"34)) ! "#$%&%'()*+),-()./ *$&-"! ! "#$%&#! '#()*+,-+.#! "! / # ! "#$%&'&()"*+,%-+.&/ 0#$,%(01"%2/ ,3+42516+7)"%"/ +8",029-2+ : ); 2&/ *<9='9%*>2/ 9? "+@A+@BCC+7)"%"/ +8",029-2A+D,02/ )9+ ! ! ! 01! 234! 215)456)! 7.04! 28)918: ; ) ! 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C<DEA! 2=? ? ! ! / "&.&V$KMN! $ News 52 Tribologie + Schmierungstechnik · volume 71 · issue 3/ 2024 BOOK RECOMMENDATION expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 \ 72070 Tübingen \ Germany \ Tel. +49 (0)7071 97 97 0 \ info@narr.de \ www.narr.de In addition to the indisputably necessary electrification of the transport sector, which is currently being ramped up, internal combustion engines will still be urgently needed in the future. Otherwise, the demand for mobility in the on-road, off-road and non-road sectors cannot be met. There is no doubt that these internal combustion engines will have to be improved regarding efficiency plus lower emissions and nowadays more and more important upgraded for zero and low carbon fuels. Even though Spark Ignition (SI) engines have been around for more than a century, there is still a lot of room for improvement, particularly in terms of power density, ignition, combustion control, and preventing uncontrolled combustion. To offer all interested developers an inspiring exchange platform for the latest developments, IAV established two exciting conferences more than two decades ago, which are now held under the heading “Two Conferences - One Goal”. This volume brings together the contributions to this conference. Content Ignition and inflammation of conventional and alternative fuels such as hydrogen, ammonia, methanol etc. - Combustion processes for alternative fuels - Prevention of irregular combustion phenomena when using conventional and alternative fuels - Methods for measurement and analysis of irregular combustion phenomena - Modern virtual development methods - Control, regulation and latest function algorithms Marc Sens (ED.) 6th International Conference on Ignition Systems for SI Engines 7th International Conference on Knocking in SI Engines 1st edition 2024, 386 p. €[D] 189,00 ISBN 978-3-381-12991-1 eISBN 978-3-381-12992-8 Checklist Author information Corresponding author: F Mailing address F Telephone and fax number F eMail All authors: F Academic titles F Full name F ORCID (optional) F Research instititute / company F Location and zip code Length F Approximately: 3,500 words Data F Word and pdf documents (both with images + captions) F Additionally, please send images as tif or jpg / 300 dpi / Please send vector data as eps Manuscript F Short and concise title F Keywords: 6-8 terms F Abstract (100 words) F Numbered pictures/ diagrams/ tables (please refer to the numbers in your text) F List of works cited After the typesetting is completed, you will receive the proofs, which you are requested to review and then give your approval to start the printing process. 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You can obtain the full open access service for a one-off article processing charge of € 1,850.00 (plus VAT). Editor in chief Dr. Manfred Jungk eMail: manfred.jungk@mj-tribology.com Publisher expert verlag Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 D-72070 Tübingen Tel.: +49 (0)7071 97 556 0 eMail: info@verlag.expert www.expertverlag.de Editor Patrick Sorg eMail: sorg@verlag.expert Tel.: +49 (0)7071 97 556 57 Tribologie und Schmierungstechnik Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie | An Official Journal of Österreichische Tribologische Gesellschaft | An Official Journal of Swiss Tribology We’re looking forward to your contribution! ISSN 0724-3472 Science and Research www.expertverlag.de Jaacob Vorgerd, Mathis Steinrötter, Alexander Thomas, Manuel Oehler Efficiency of high speed spur gears with an isotropic superfinishing Markus Grebe, Henrik Buse, Alexander Widmann Comparison of different standard test methods for the evaluation of greases for rolling bearings under small oscillating movements Matthias Kröger, Jim Gerschler, Ringo Nepp, Christian Berndt Rotary shaft seals at high temperatures Dirk-Olaf Leimann Thoughts on a standardized FE-8 Test for the assessment of the WEC-carrying capacity of lubricants in rolling bearings Michael Hilden, Gerd Dornhöfer, Harald Dietl Improvements in brake fluid standardization to avoid noise & wear problems