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Tribologie und Schmierungstechnik EDITOR IN CHIEF MANFRED JUNGK 1 _ 26 VOLUME 73 Tribology—Lubrication Friction Wear An Official Journal of Gesellschaft für Tribologie An Official Journal of Österreichische Tribologische Gesellschaft An Official Journal of Swiss Tribology Issue 1 | 2026 Volume 73 Editor in chief: Dr. Manfred Jungk Tel.: +49 (0)177 1902330 eMail: jungk@verlag.expert www.mj-tribology.com Editorial director: Ulrich Sandten-Ma Tel.: +49 (0)7071 97 556 56 / eMail: sandten@verlag.expert Editor: Dr. rer. nat. Erich Santner Tel.: +49 (0)2289 616136 / eMail: esantner@arcor.de Editorial office: Angelika Böer Tel.: +49 (0)7071 97 556 57 / eMail: boeer@narr.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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Receiving the journal for a reduced price obligates the subscriber to purchase the whole volume. If the subscription is terminated prematurely, the unit price will be charged. Higher power cancels delivery obligation. Place of performance and jurisdiction: Tübingen. ISSN 0724-3472 ISBN 978-3-381-15851-5 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 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0001 The first Report was issued in 2014 followed by reports every 3 years. While overall megatrends after the turn of the millennium were Water, Aging Population, Urbanification, Climate etc. STLE took a look closer to Tribology related topics. 2014 the topics surveyed included Outlook on Field, Energy, Manufacturing, Transportation, Atomistic Processes, Government Regulation, Safety and the Environment and Basic Human Needs. Top ranked issues were ability to attract talent, decreasing use of fossil fuels, increased productivity and efficiency, high efficient vehicles, increasing need for higher education, more restrictive emission standards, particulate matter emissions and reuse of water after it’s treated. 2017 the topics discussed included Transportation, Medical Health, Energy, Manufacturing, Communications and Data Management, Atomistic Processes, Outlook on field, Government Regulation, Safety and the Environment and Basic Human Needs. Lower viscosity motor oils, Soft matter, Natural gas, Reduction in costs of production, Continuous monitoring of lubricants, Investigating formation of films on surfaces, The ability of the field to attract talented and educated employees, Industrial waste regulations, Automation replacing workers and Production of sustainable energy were mentioned as key issues in each category. 2020 the focus lay on Transportation, Energy, Manufacturing, Medical/ Health, Workforce Issues, Research Funding, Materials Cost and Availability, Safety and the Environment, Basic Human Needs, Communications and Data Management and Government Regulation. The top issues mentioned were EV Transmission, Energy Storage, Increased Productivity, Prosthetics, Aging Workforce, Batteries, Security of the supply chain, Sustainable Energy and Raw Materials Regulations. 2023 key findings related to each of the topics Supply Chain, Sustainability, Electrification, Manufacturing, Medical/ Health and Government Regulation were Preparation and distribution of lubricants are very complex, Motivation is high for leading academia, governments, industry and research laboratories to minimize the negative consequences of global warming, Rapid increase in demand for digital content, Growing use of robots/ cobots, Growing use of medical implants is leading to the greater use of tribology and More chemical substances are under scrutiny. 2026 the four leading trends explored are AI, Manufacturing, Thermal Management and Decarbonization. A lack of unified and coalesced data with standardized tests slow down it’s flow. Deployment of better lubricants utilizing digitalization will increase efficiency in manufacturing facilities. Lubricant technologies are aimed at dissipating heat in electric vehicles to the operation of data centers. Incentives moving away from fossil fuels have reduced and slowed down the development of alternatives. While transportation developed from using low viscosity oils to thermal management over the 12-year span, manufacturing issues centered around efficiency and supply. Decarbonization and Big Data will keep us busy to explore that Tribology is everywhere. Your editor in chief Manfred Jungk STLE publishes 5 th Report on Emerging Issues and Trends in Tribology and Lubrication Engineering Events 2 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 Events Date Place Event 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 ► 02.09. - 04.09.26 Valpre, France 51 th Leeds Lyon Symposium ► 08.09. - 11.09.26 Gdansk, Poland 44 th Polish Tribology Conference ► 15.09. - 17.09.26 Düsseldorf, Germany Lubricant Expo Europe ► 20.09. - 25.09.26 Rio de Janeiro, Brazil 8 th World Tribology Congress ► 28.09. - 30.09.26 Wernigerode, Germany 67 th Annual GfT Conference ► 30.09. - 02.10.26 Amsterdam, The Netherlands 61 st UEIL Annual Congress ► 09.11. - 12.11.26 Hoofdorp, The Netherlands ELGI Autumn Events ► 18.11. - 19.11.26 Fellbach, Germany 33. Kühlschmierstoffforum 2026 ► 19.11. - 21.11.26 Kaunas, Lithuania 13 th International Tribological Conference BALTTRIB 2026 ► 23.11. - 24.11.26 Lund, Sweden 6 th International conference on polymer tribology (PolyTrib 2026) Contents 3 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 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 73, Issue 1 July 2026 5 Jakob Gleichner, Markus Golek Development of a squeeze film damper test rig - comparison of measurement and simulation results 14 Felix Leuf, Georg Jacobs, Jan Euler, Patrick Laubichler, Michael Mirza, Thomas Decker, Julian Röder Evaluation of the start-up behavior of a novel spherical plain bearing as a main bearing for wind turbines based on scaled tests 22 Yongzhen Lin, Matthias Kröger, Jenny Köckritz, Deeksha Lohani, Ringo Nepp Tribological Investigations of Radial Shaft Seal with Focus on 3D Printed Counter Surface 32 Marius Hofmeister, Katharina Schmitz Increasing efficiency of diesel injection pumps by targeted blending of bio-hybrid fuels 1 Editorial STLE publishes 5th Report on Emerging Issues and Trends in Tribology and Lubrication Engineering 2 Events Science and Research Preface For authors Authors of scientific contributions are requested to submit their manuscripts directly to the editor, Dr. Jungk (see inside back cover for formatting guidelines). Anzeige 4 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 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 45 % 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 45 % subsidy possible Further information and registration at www.tae.de/ weiterbildung/ tribologie-reibung-verschleiss-schmierung/ Introduction Squeeze film dampers (SFDs) have been used for many decades to effectively reduce vibrations in the bearings of rotating engine components. In combination with a rolling bearing, as shown schematically in Figure 1, a thin lubricating film separates the outer ring of the rolling bearing, which is structurally prevented from rotating, from the surrounding housing. Depending on the design, the outer ring can be guided in a centralized or non-centralized manner; in centralized variants, a squirrel cage usually provides the centering restoring force [1]. If the outer ring moves relative to the housing, hydrodynamic pressure is generated in the lubricating film, which absorbs energy from the vibration and thus provides effective damping. In this way, critical resonances can be reduced, amplitudes limited, and the operational stability of the rotor system increased [13]. Despite widespread use and extensive research, questions remain regarding the precise physical description and robust modeling under realistic conditions. Key challenges include nonlinear effects of the lubricating film, cavitation phenomena, thermal couplings between fluid and structure, and influences from eccentricity, rotor misalignment, and manufacturing tolerances [15]. The interplay of damping and stiffness, especially in the presence of circumferential grooves or piston rings, also requires valid models that reliably represent dynamic behavior over a wide operating range [14]. In order to systematically investigate these issues, a specially designed test rig was developed at the University of Kassel that enables experiments with controlled shaft movement. The modular design allows the use of different SFD geometries, variable lubricating film thicknes- Science and Research 5 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 Development of a squeeze film damper test rig - comparison of measurement and simulation results Jakob Gleichner, Markus Golek* submitted: 26.09.2025 accepted: 18.12.2025 (peer review) Presented at GfT Conference 2025 At the University of Kassel, a test rig for squeeze film dampers was developed to enable realistic experiments with controlled orbital motion. Using this setup, measured data were obtained and subsequently used to validate FEM-based simulation software designed to analyze both hydrodynamics and heat transfer in squeeze film dampers. Among other things the validation process compared hydrodynamic forces and pressure obtained from experiments with corresponding simulation results. Two different squeeze film damper configurations were examined to assess accuracy and reliability. The results demonstrate that the developed rig and validated simulations provide valuable insights into squeeze film damper behavior under realistic operating conditions. Keywords squeeze film damper, test rig, reynolds equation, thermo-hydrodynamic, inertia effects, rotordynamics Abstract * Jakob Gleichner, M. Sc. Orcid-ID: https: / / orcid.org/ 0009-0003-6148-2138 Markus Golek, M. Sc. Orcid-ID: https: / / orcid.org/ 0009-0004-1900-6290 Institut für Antriebs- und Fahrzeugtechnik - Lehrstuhl Maschinenelemente und Tribologie - Universität Kassel Universität Kassel Figure 1: Schematic structure of an SFD cords, among other things, displacement and force variables, pressure and temperature fields in the lubricating film, as well as the volume flow and accelerations. This is supplemented by flexible temperature control of the housing and the lubricant, so that thermally coupled effects can also be mapped in stationary and transient operating conditions. The structure, operating principle, and mode of operation of the test rig are systematically described below. The aim is to quantitatively characterize the hydrodynamic and thermal processes in squeeze film dampers, identify nonlinear effects, and validate the physical models implemented in the institute’s own simulation software with reliable test data. The test rig is shown in Figure 2 and serves as a reference for the assemblies and measurement positions explained in the text. Beyond pure validation, the test rig enables sensitivity and parameter variation studies (e.g., influence of eccentricity, lubricant temperature, speed, and sealing concept), the derivation of practical design guidelines, and the comparison of different damper configurations under identical boundary conditions. The system thus contributes to the reliable design of SFD systems and to the transferability of simulation results to real applications in the rotor bearings of drive systems. The test shaft is actuated by two electromagnetic shakers oriented orthogonally. Analogous to the San Andrés test rig, the circular or elliptical motion is created by coordinated control of both actuators with defined amplitude and phase angle [2]. In this way, circular or elliptical paths can be realized around a selectable reference point in the housing bore - both centrally and specifically eccentrically. The movement is controlled in a closed loop so that the path shape, frequency, and amplitudes are precisely maintained. Science and Research 6 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 ses, and configurable circumferential grooves, as well as the use of piston ring seals to specifically influence leakage and pressure distribution. This makes it possible to investigate almost all operating conditions relevant to aircraft engines, from stationary to transient load cases, including defined speed and temperature ranges as well as feed pressures. The test rig is designed to generate reproducible, highresolution measurement data. This includes, for example, the shaft’s orbit curves, pressure and temperature fields in the lubricating film, and force measurements on the shaft. The data sets obtained in this way form the basis for comprehensive validation of specially developed simulation software that takes into account both hydrodynamics and heat transfer in squeeze film dampers. Based on this validation, model parameters can be calibrated and boundary conditions verified so that the simulation can be reliably used for design studies, sensitivity analyses, and design-related optimizations. All in all, the test rig at the University of Kassel bridges the gap between theory and application: it enables the reproducible simulation of complex operating conditions, provides reliable measurement variables for comparison with numerical models, and supports the development of high-performance, low vibration bearing concepts for modern aircraft engines. The insights gained contribute to a more precise understanding of the physical mechanisms in SFDs, reduce design uncertainties, and further increase the reliability of rotordynamic systems in real-world operation. Test rig A modular test rig for the experimental investigation of squeeze film dampers was developed at the Chair of Machine Elements and Tribology at the Institute for Drive and Vehicle Technology (iaf-mt) at the University of Kassel. The development of the system encompassed the entire process chain - from conceptual requirements analysis and preliminary design dimensioning to numerical analyses at component and system level, detailed design, manufacturing, assembly, and commissioning. Particular attention was paid to design-oriented parameterization in order to be able to vary geometries, sealing concepts, and operating variables in a reproducible manner. The core of the test rig is a precisely controlled shaft movement that can be used to specify defined excitation curves in terms of amplitude, frequency, and phase angle. The test rig allows the use of different SFD geometries, variable lubricating film thicknesses, and circumferential grooves, as well as operation with piston ring seals for targeted influence on leakage and pressure distribution. The measuring chain is multi-channel and re- Figure 2: SFD test rig The design is modular so that squeeze film damper of different geometries and configurations can be investigated. This includes variable gap heights, oil supplies via bores with or without circumferential grooves, and sealing concepts with piston rings for targeted influence of leakage and pressure distribution. Interchangeable inserts allow quick modifications without changing the rest of the measurement technology. Two geometrically similar bearing sizes were provided for the investigation (Figure 3). The smaller bearing corresponds to a 2/ 3 scale of the large bearing. This choice of scale allows the operating conditions defined in the test rig, in particular the load and speed ranges, to be run through completely and reproducibly. Both variants are identical in terms of bearing design, material, lubrication, and boundary conditions. The small design reduces the required shaker power without fundamentally changing the contact and kinematic conditions. This allows a comparison of the measured variables between the two bearing sizes and the transferability of the results to be evaluated. Depending on the moving mass, the available working range covers excitation frequencies up to 500 Hz. Relative eccentricities up to ε = 0.9 can be represented. The practically achievable combinations of amplitude and frequency are limited by the maximum force of the shakers. This allows both small linearized and highly nonlinear operating states to be specifically addressed. Oil conditioning with heat exchangers ensures constant and reproducible lubrication conditions, enabling oil outlet temperatures of up to approx. 120 °C while also providing stable high-volume flows. The inlet pressure is variably adjusted between one and five bar via an external pump. The measurement technology is synchronized across multiple channels. The forces are recorded using force measurement flanges equipped with strain gauges (Figure 4). Science and Research 7 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 Figure 4: Load cell assembly Figure 3: Comparison of the two bearing sizes that can be tested on the test rig. left: small bearing (2/ 3 scale of the large bearing), right: large bearing Data measurement is multi-channel and time synchronized. The forces are measured using strain gauge-based load cells, which are arranged in the force flow between the shaker and the shaft. The force is transmitted via spring plates, so that the forces initiated by the shakers are applied to the shaft in a defined manner and measured precisely. In addition, seven acceleration sensors on the shaft and housing record the dynamic conditions. The inertial components of the moving components are determined from the measured accelerations and subtracted from the total forces in order to evaluate the hydrodynamic forces in isolation. To determine the hydrodynamic forces, the measured total force is first filtered appropriately. The acceleration peratures, the material properties (conductivity, density, specific heat capacity), and the boundary conditions (convection, heat input from oil) are incorporated into a numerical model. In the next step, the thermoelastic deformation of the housing is calculated from the determined temperature distribution. For this purpose, the thermal expansions and component stiffnesses are considered in a coupled structural calculation. The result is a deformed housing profile under operating heat, the so-called warm contour, see Figure 7. The operating-temperature contour of the bearing is crucial for correctly assessing the clearance, film load-carrying capacity, and load distribution under real operating conditions. By combining dense temperature measurement, inverse thermal analysis, and subsequent structu- Science and Research 8 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 measured at the rotor is then linked to the known mass properties in order to calculate the corresponding inertial force. The difference between the filtered total force and the inertial force provides the hydrodynamic force for each axis (see Figure 5, blue). The pure inertial forces were experimentally validated in a separate reference setup without housing and without oil. A closely spaced network of 32 temperature sensors on the surfaces of the housing and the shaft provides information on the spatial and temporal temperature distribution in the solids. The measuring points shown in Figure 6 record the local temperature on the housing and on the shaft. Based on this data, the volume temperature distribution in the housing and shaft is reconstructed using an inverse heat conduction calculation [8]. To do this, the measured tem- Figure 5: Inertial forces Figure 6: Temperature measuring points on the small housing ral calculation, temperature-related housing deformation can be quantified and incorporated into the validation of the simulation models. In addition, uncertainties about the alignment curves are estimated to reliably evaluate the quality of the reconstructed thermally expanded bearing geometry. Each test is based on a defined alignment curve of the shaft. For this purpose, the shaft is guided along the housing contour. This procedure creates reproducible boundary conditions, enables a positionresolved comparison of measurement and simulation, and improves the traceability of the determined hot contour over the entire circumference. With a special shaft configuration, it is also possible to measure the pressure directly in the lubricating film and to record the oil temperature in the gap directly. The oil supply is continuously monitored by additional pressure sensors, a volume flow sensor, and additional temperature sensors. Various operating modes are also available: stationary centric or eccentric circular paths and transient tests. This combination of flexible excitation, stable boundary conditions, and high-resolution measurement chain provides a robust basis for quantitatively describing hydrodynamic and thermal effects in squeeze film damper, validating models, and systematically investigating sensitivities to geometry and operating parameters. Simulation The most important requirement for the simulation software is the realistic representation of physical effects with short calculation times to quickly obtain information about system behavior within the scope of rotor dynamics simulations. The crucial part of the model is the hydrodynamics, which represent the pressure distribution and thus the reaction forces of the squeeze film damper. In addition, the hydrodynamics are influenced by thermal effects. Both parts of the model are described in detail below. The basis for hydrodynamics is the Navier-Stokes equations for incompressible fluids (1) [7]. These equations balance the forces in volume elements. (1) In addition to the velocity vector u = {u, v, w} T and the pressure p, the material parameters viscosity (η) and density (ρ) are the decisive influences. External forces acting on the mass of the volume element are denoted by f = {f x , f y , f z } T . A special feature of the squeeze film damper geometry is the very small bearing clearance compared to the radius. This fact allows for a number of simplifications to the model. Due to the small ratio of the gap height to the circumference and width of the squeeze film damper and with the recommended edge length ratios of the three-dimensional elements (cf. [3]) a very fine mesh is generated, which results in long calculation times. It is therefore necessary to simplify the three-dimensional model by transforming it into a two-dimensional model. Due to the size ratio of bearing clearance to circumference and width, the following simplifications are permissible. The external forces from equation (1) can be neglected. The pressure is constant across the gap height. In addition, the velocities in the gap height direction and the velocity gradients in the circumferential and width directions are negligible [6]. Taking these simplifications into account, as well as the variable viscosity across the gap height (viscosity depends on a parabolic temperature distribution) we arrive at Dowson’s generalized Reynolds equation (2) [10]. Since the inertial forces cannot be neglected in squeeze film dampers, the generalized Reynolds equation is extended by G 1 and G 2 [4]. (2) Here, h is the gap height and u the velocity of outer ring’s surface. F 1 and F 2 are combinations of the Dowson integrals that integrate viscosity over the gap height (cf. (3)) [10]. (3) The change in density over time is adjusted in accordance with the cavitation model developed by Kumar and Booker to ensure that mass conservation is maintained. The exact implementation is described in [12]. � + ⋅ � ����������� = − � + 2 ��� + ⏟ � 1 � = ( 2 ) + ℎ + ℎ + 1 ( 0 ) + 2 ( 0 ) 1 = � 2 − 2 � ℎ 0 , 2 = ∫ ℎ 0 ∫ 1 ℎ 0 ℎ 0 Science and Research 9 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 Figure 7: Inverse temperature distribution of the small housing (see measured points in Figure 6) (8) (9) The finite element method is suitable for numerical implementation. The weak formulation for hydrodynamics is given in equation (10) [5][8]. The exact derivation is described in [12]. (10) In addition to the lubricating film, the oil supply flows are part of the modelling, so that no pressure boundary conditions need to be set in the lubricating film. The pipe flow is derived from the Hagen Poiseuille equation for 1D elements with flow direction in x and added to the stiffness matrix of the squeeze film. Both equations are coupled via the flow balancing [12]. (11) The weak formulation of the energy equation for fluids is given in equation (12)[5]. (12) The weighting function W i consists of the shape function ψ i and a perturbation term P i that shifts the integration points upstream to prevent oscillations in the solution. The energy equation for fluid is coupled with the energy equation for solids in a system of equations. The surface temperatures of the rolling bearing T J and the housing T C become unknowns, and the system matrix of the lubricating film is integrated into the system matrix of the solid [12]. Results and Discussion The software is validated using the measurement data from the test rig described above. The focus here is on comparing the hydrodynamic forces. The hydrodynamic pressures in the lubricating film are also compared. The input for the simulation consists of the measured distances, oil supply temperatures and volume flows. One second of a measurement (M) was evaluated stochastically with the specified parameters frequency, re- ̅ = − 1 ℎ � = 2 ∫ 1 ℎ 0 + 1 � � 2 + 1 � � 2 � Ω 1 �∂ψ ∂ ∂ψ ∂ + ∂ψ ∂ ∂ψ ∂ � Ω ⋅ = � � 2 ∂ψ ∂ + ℎ ∂ ∂ ψ + ∂ℎ ∂ ψ + Ω 1 ( 0 ) ψ + 2 ( 0 ) ψ � Ω + � Γ ̇ Γ ψ � 4 8 ⋅ = � ̇ � � �ℎ � � + ̅ � + ℎ 12 � + Ω ℎ � + �� Ω ⋅ � = � � � � + 6 ℎ � + ��� Ω Ω Science and Research 10 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 The terms G 1 and G 2 result from the derivation of the inertia effects for the two-dimensional application case according to Hamzehlouia [4]. Inertia can no longer be neglected for higher amplitudes, excitation frequencies and eccentric orbit centres [11]. A perturbation calculation is used to approximate the inertia effects in the lubricating film, assuming that the inertia of the fluid has no influence on the fluid velocities. G 1 in equation (4) are the temporal inertia effects and G 2 in equation (5) are the convective inertia effects [4]. (4) (5) Pressure p 0 is obtained by solving the Reynolds equation without G 1 and G 2 . The pressure distribution p 0 is the basis to calculate G 1 and G 2 and solving the extended Reynolds equation. The thermal effects are described by the energy equation for fluids averaged over the gap height and solved on the same two-dimensional mesh as the hydrodynamics in order to efficiently couple both equations. The energy equation for fluids is also coupled with the energy equation for solids in one system of equations and captures the heat exchange between the lubricating film and surrounding structures. The 2D energy equation for fluids is given in equation (6) [5]. (6) Here, u̅ and v̅ are the fluid velocities averaged over the gap height, which can be determined using the pressure distribution (see equations (7) and (8)). On the righthand side of the equation is the hydrodynamic dissipation averaged over the gap height, which is also derived from the pressure distribution and describes the heat introduced into the system by friction (see equation (9)). The Dowson integrals are also necessary to determine these variables. (7) 1 ( 0 ) = 2 12η � h 2 ∂ 2 h ∂ t 2 + h 6η ∂ h ∂ x � h 3 ∂ 2 p 0 ∂ t ∂ x + 3h 2 ∂ h ∂ t ∂ p 0 ∂ x �� ℎρ � � ∂ ∂ + ̅∂ ∂ � − ∂ ∂ �ℎλ ∂ ∂ � − ∂ ∂ �ℎ � − λ �∂ ∂ � 0 ℎ = � � = − 1 ℎ + ℎ 2 − 7 ℎ 6 72 ∂ℎ ∂ ∂ 0 ∂ ∂ 2 0 ∂ 2 − ℎ 7 72 �∂ 2 0 ∂ 2 � 2 − ℎ 7 72 ∂ 0 ∂ ∂ 3 0 ∂ 3 − 7 ℎ 6 144 ∂ℎ ∂ ∂ 2 0 ∂ ∂ ∂ 0 ∂ − ℎ 7 144 ∂ 3 0 ∂ 2 ∂ ∂ 0 ∂ − ℎ 7 144 � ∂ 2 0 ∂ ∂ � 2 − 7 ℎ 6 144 ∂ℎ ∂ ∂ 0 ∂ ∂ 2 0 ∂ 2 − ℎ 7 144 ∂ 2 0 ∂ 2 ∂ 2 0 ∂ 2 − ℎ 7 144 ∂ 0 ∂ ∂ 3 0 ∂ ∂ 2 − 5 ℎ 6 144 ∂ℎ ∂ ∂ 2 0 ∂ ∂ ∂ 0 ∂ − 5 ℎ 6 144 ∂ℎ ∂ ∂ 0 ∂ ∂ 2 0 ∂ 2 − ℎ 7 144 ∂ 3 0 ∂ ∂ 2 ∂ 0 ∂ − ℎ 7 144 � ∂ 2 0 ∂ ∂ � 2 − ℎ 7 144 ∂ 2 0 ∂ 2 ∂ 2 0 ∂ 2 − ℎ 7 144 ∂ 0 ∂ ∂ 3 0 ∂ 2 ∂ − ℎ 7 72 �∂ 2 0 ∂ 2 � 2 − ℎ 7 72 ∂ 0 ∂ 3 0 ∂ 3 � − 5ℎ 6 144 ∂ 2 ℎ 2 �∂ 0 ∂ � 2 − 5ℎ 6 72 ∂ℎ ∂ ∂ 0 ∂ 2 0 ∂ 2 2 2 ( 0 ) = ρ 2 12 η 3 � − 5 ℎ 5 20 �∂ℎ ∂ � 2 �∂ 0 ∂ � 2 lative eccentricity ε and oil temperature remaining constant. The double interval of the standard deviation σ, including the measurement error F err , around the mean value μ is shown. The simulation is marked with S. Starting with the small squeeze film damper configuration Figure 8 shows good correlation between measurement and simulation for medium frequency and low eccentricity at low oil temperature. Phase and amplitude are very similar. However, small peaks in the force curve at the feed holes are only visible in the simulation. Figure 9 shows a similar picture with good agreement between measurements and simulation for high frequencies. The forces are correspondingly greater. For the large squeeze film damper configuration, there is very good agreement at low frequency and medium relative eccentricity at oil temperature of 50 °C in Figure 10. The simulation of the large configuration does not show any peaks at the oil supply holes. With the same frequency and oil temperature Figure 11 shows the comparison for high relative eccentricity. There is a moderate deviation between the simulation and the measurement. However, in terms of quality, the results match up well. The deviation is due to the accuracy of the squeeze film damper clearance determination. Due to the influence of temperature, the clearance differs from the nominal value and is not known exactly. Calculating thermal expansion using inversely determined temperature fields helps to determine the warm clearance. However, at high eccentricity, every micrometer has a strong influence on the forces. Figure 12 shows a similar picture for high eccentricity at higher frequencies and high oil temperatures. Here too, there are moderate deviations, which can be attributed to the inaccuracy of the clearance determination for warm components. However, given these inaccuracies, the deviations are acceptable. The forces are an integral quantity and do not show the details of the hydrodynamics. For local validation, three pressure sensors are installed that measure the pressure directly in the lubricating film. The sensors are located at the circumferential angles 1/ 2 π, π and 3/ 2 π on the same axial position at B/ 4 (B is the width of the squeeze film damper). With a cylindrical clearance, approximately equal amplitudes can therefore be expected for an orbital motion around the centerline. Figure 13 shows the pressure over the phase angle for the same test as in Science and Research 11 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 Figure 8: Comparison at 200 Hz, ε = 0.3, 40 °C Figure 9: Comparison at 333 Hz, ε = 0.3, 40 °C Figure 10: Comparison at 66 Hz, ε = 0.5, 50 °C Figure 11: Comparison at 66 Hz, ε = 0.8, 50 °C Figure 12: Comparison at 200 Hz, ε = 0.8, 100 °C Comparisons of measurements and simulations show good agreement over the entire operating range. At high relative eccentricities, the deviations become slightly larger, which is due to uncertainties in the determination of warm clearance and the sensitivity of hydrodynamic forces to small differences in the gap at high eccentricities. The software with the described model is therefore a suitable tool for determining the influence of squeeze film dampers on rotor dynamic systems. The comparisons for an open-end squeeze film damper without circumferential groove provide a good basis for validation with circumferential grooves and sealings. Literature [1] Chatzisavvas, I.; Arsenyev, I.; Grahnert, R. Design and optimization of squirrel cage geometries in aircraft engines toward robust whole engine dynamics. Applied and Computational Mechanics 2023, 17, 93-104. [2] San Andrés, L. Force coefficients for a large clearance open ends squeeze film damper with a central feed groove: Experiments and predictions. Tribology International 2014, 71, 17-25. https: / / doi.org/ https: / / doi.org/ 10.1016/ j.triboint.2013.10. 021 [3] M. Schmidt, P. Reinke, A. Rabanizada, S. Umbach, A. Rienäcker, D. Branciforti, U. Philipp, M. Bargende, A.-C. Preuß, K. Pryymak, and G. Matz. Numerical study of the three-dimensional oil flow inside a wrist pin journal. Tribology Transactions, 63(3): 415-424, 2020 [4] S. Hamzehlouia. Squeeze Film Dampers in High-Speed Turbomachinery: Fluid Inertia Effects, Rotordynamics, and Thermohydrodynamics. PhD thesis, Universtiy of Toronto, Mechanical and Industrial Engineering, 2017 [5] D. Jaitner. Effiziente Finite-Elemente-Lösung der Energiegleichung zur thermischen Berechnung tribologischer Kontakte. PhD thesis, Universität Kassel, 2017. [6] O. R. Lang and W. Steinhilper. Gleitlager. Springer Berlin, Heidelberg, 1 edition, 1978 [7] H. Oertel. Strömungsmechanik. Methoden und Phänomene. Originalveröffentl. im Springer-Verl., Berlin, 1995. Universitätsverlag Karlsruhe, 2005 [8] A. Rienäcker. Instationäre Elastohydrodynamik von Gleitlagern mit rauhen Oberflächen und inverse Bestimmung der Warmkonturen. PhD thesis, RWTH Aachen, 1995 [9] A. Kumar and J. F. Booker. A Finite Element Cavitation Algorithm. Journal of Tribology, 113(2): 276-284, 04 1991 [10] D. Dowson. A generalized reynolds equation for fluidfilm lubrication. International Journal of Mechanical Sciences, 4(2): 159-170, 1962 [11] L. San Andrés. Modern Lubrication Theory. Notes 13: Squeeze film dampers: operation,modelsandtechnicalissues.TexasA&MUniversityDigitalLibraries, 2010. http: / / oaktrust.library.tamu.edu/ handle/ 1969.1/ 93197 Accessed: 2024-07-02 Science and Research 12 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 Figure 10. The pressures P 2 and P 3 match well in both amplitude and phase, P 1 has a small deviation in phase. All three positions have similar amplitude. For high relative eccentricity at low frequencies (see Figure 14), the pressure values also match very well, even though there is a slight phase shift. The different amplitudes are due to a tilt of the shaft in the casing. Overall, measurements and simulations show good agreement for the two configurations, the small and large squeeze film damper. Figure 13: Pressure at 66 Hz, ε = 0.5, 50 °C Figure 14: Pressure at 66 Hz, ε = 0.8, 50 °C Conclusions A squeeze film damper test rig was developed and put into operation to analyze squeeze film dampers under real operating conditions in various measurement series and to provide data for validating the developed software. The thermo-hydrodynamic simulation model was converted to two dimensions, which greatly reduced the calculation times. For hydrodynamics, this results in an extended form of Reynolds equation, which, in comparison to the generally known form, takes into account local and convective inertia effects, mass-conserving cavitation and variable viscosity due to the temperature distribution. The thermal effects are included in full coupling of the fluid with the surrounding components. [12] M. Golek. Thermo-hydrodynamische Modellierung und Validierung von Quetschfilmdämpfern für rotordynamische Anwendungen, PhD thesis in preparation, Universität Kassel [13] San Andrés, Luis; Koo, Bonjin; Jeung, Sung-Hwa. (2018). Experimental Force Coefficients for Two Sealed Ends Squeeze Film Dampers (Piston Rings and O-Rings): An Assessment of Their Similarities and Differences. Journal of Engineering for Gas Turbines and Power, 141(2), 021024-.doi: 10.1115/ 1.4040902, p.2, 1-4 [14] San André s, Luis. (2012). Damping and Inertia Coefficients for Two Open Ends Squeeze Film Dampers With a Central Groove: Measurements and Predictions. Journal of Engineering for Gas Turbines and Power, 134(10), 102506-.doi: 10.1115/ 1.4007058 [15] Edoardo Gheller, Nicolas Grigat, Adolfo Delgado & Paolo Pennacchi (07 Jul 2025): Multifrequency Response of a Squeeze Film Damper with Incipient Air Ingestion, Tribology Transactions, DOI: 10.1080/ 10402004.2025.2514760, p.5 Science and Research 13 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0002 Science and Research 14 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0003 1 Introduction Wind energy is particularly important for a sustainable energy supply. In Germany, wind energy represents the most significant source of electricity [1]. Still, to spur on the set up of even more wind turbines (WT) the economic efficiency needs to be improved. An increase of the economic efficiency can be achieved by reducing the levelized cost of electricity (LCOE). This can be done by reducing operating and maintenance costs, which can be realized by reducing component failures. Research indicates that a substantial portion of WTs experience failures due to damage to the main bearings before reaching their intended design lifetime of up to 25 years [2]. The main bearings support the weight of the rotor and bear the loads induced by the wind. Typically, rolling element bearings are used as WT main bearings. When such a main bearing experiences a failure, especially in offshore wind turbines, the repair costs can be high - specialized crane ships costing over € 200,000 per day are required for a main bearing replacement [3]. One option to reduce repair costs is to use a segmented plain bearing as the WT main bearing. Segmented plain bearings allow for individual segments to be replaced at significantly lower costs directly up-tower. In recent years, various concepts for plain bearing main bearings have been developed. While there are differences in functio- Evaluation of the start-up behavior of a novel spherical plain bearing as a main bearing for wind turbines based on scaled tests Felix Leuf, Georg Jacobs, Jan Euler, Patrick Laubichler, Michael Mirza, Thomas Decker, Julian Röder* submitted: 24.09.2025 accepted: 6.01.2026 (peer review) Presented at GfT Conference 2025 Wind energy is an ever-increasing part of the German and European energy supply. A high reliability and low operational cost are crucial for the economic viability of wind turbines (WTs). One component that often fails before reaching its specified service life is the WTs main bearing. At present, only rolling element bearings are used as main bearings in series production. Segmented plain bearings potentially offer drastic improvements regarding WT downtime in case of failure, as their segments can be exchanged up-tower without the need to dismount the rotor. Plain bearings as WT main bearings are therefore subject to public and industrial research. However, plain bearings experience a high breakaway torque during the WTs start up and high friction during idling operation. The Miba SJB™ is a currently researched plain bearing solution that is characterized by a spherical sliding surface. In this work test rig results for this plain bearing concept are presented. The bearing was tested as a scaled prototype with accordingly downscaled loads on a test rig able to produce forces and bending moments corresponding to loads from a 1 MW WT. Main focus of this work is the analysis of the bearing’s behavior under start-up and idling conditions. After long downtimes, a favorable behavior of the breakaway torque during start-up is determined and a low transition sliding speed is measured during idling operation. The test results demonstrate that WTs with the tested bearing as main bearing can also start up at low wind speeds, as hydrodynamic operation occurs even at low sliding speeds. Keywords segmented plain bearings, main bearings, wind turbines, idling, start-stop cycles Abstract * Felix Leuf, M.Sc. 1 (corresponding author) Prof. Dr.-Ing. Georg Jacobs 1 Jan Euler, M.Sc. 1 Patrick Laubichler, M.Sc. 2 Michael Mirza, M.Sc. 2 Thomas Decker, M.Sc. 1 Julian Röder, M.Sc. 1 1 Chair for Wind Power Drives, RWTH Aachen University Campus-Boulevard 61, 52074 Aachen 2 Miba Gleitlager Austria GmbH Dr.-Mitterbauer-Str.3, 4663 Laakirchen nality among these designs, they generally consist of individual sliding segments that facilitate easy replacement. In addition to the geometry of the bearings’ sliding surfaces, the concepts differ significantly in that the sliding segments can be attached either to the rotating shaft [4] or to the stationary outer ring [5, 6]. The subject of this paper is a spherical plain bearing which was developed for three-point suspension drivetrains. Due to its spherical geometry, it can support both axial and radial forces. The outer sliding surface remains continuous and stationary, whereas the inner sliding surface is segmented and rotates during operation. The segments can be accessed and removed axially through a mounting window. [4] The bearing concept under investigation is new. Thus, its behavior is largely unknown. This paper presents results from a test campaign including start-stop and idling operation that is realized on a load application system (LAS) test rig with the Miba SJB™ bearing. A downscaled version of the bearing is used so that the tests can be carried out on the aforementioned test rig. The bearing was designed for a 4.x MW WT and is geometrically scaled by a factor of 1: 3.x. The tests aim to investigate the breakaway behavior of the bearing during startup with the objective to determine its breakaway torque. Consideration is given to both breakaways after downtime under load and breakaways at different bearing temperatures. Additionally, idling tests are conducted to estimate the transition sliding speed under load conditions. The aim of this work is to investigate whether the bearing can start up at low wind speeds. 2 Test setup The tests presented are part of the “NextMBU” research project [7]. The test procedures were carried out on the LAS test rig at the Center for Wind Power Drives (CWD). The setup described below was used for this purpose. Test rig setup The test rig consists of an electric drive, the LAS, and a machine bed on which different test specimens can be installed (Figure 1). The LAS consists of four hydraulic cylinders that can apply loads in four degrees of freedom: forces in both xand z-direction as well as bending moments around both yand z-axes, as illustrated in Figure 1. Force limits range from 0 kN to 480 kN in the axial direction and -200 kN to 120 kN in the radial direction. The maximum bending moment that can be applied is 190 kNm, and the maximum torque is 18.5 kNm. The torque is measured using a torque transducer integrated into the shaft between the drive and the LAS. In addition, the temperature of the oil sump is measured. Axial and radial displacements of the shaft are measured using displacement sensors. To measure radial displacement, three distance sensors are distributed evenly around the shaft circumference next to the bearing on the side facing away from the load unit. Bearing concept The tested bearing is a spherical plain bearing. Its inner sliding surface consists of eight segments distributed around the circumference. The spherical geometry enables the bearing to support radial and axial loads. Bending moments (M y and M z ) are additionally absorbed by the WT gearbox, which is replaced for these tests by a support bearing on the side facing away from the load unit. The bearing consists of the following components: an inner ring, an outer ring, the bearing segments, a hous- Science and Research 15 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0003 Figure 1: Test rig with LAS unit tact. Therefore, the bearing experiences mixed friction during system startup [8]. To test the bearing behavior during start-up, start-stop tests are performed. To minimize statistical effects, the tests are repeated 30 times. The sliding speed increases up to 0.2 m/ s in 15 seconds under load, is held for 330 seconds and then reduced back to zero m/ s (Figure 3). Additionally, tests are performed in which the downtime under load (drainage) varies from five minutes to 17 hours. The oil temperature is kept constant at 20 °C. In addition, start-stop tests are performed at various oil temperatures: 20, 30, 40, and 50 °C. The oil temperature is controlled in the external oil circuit. The temperature of the oil in the bearing differs from this controlled temperature and is measured using sensors in the bearing’s oil sump. Another critical condition for plain bearings as main bearings in WTs is idling, during which loads are applied at low speeds. Figure 4 shows the flange loads of two different idling load cases. In both cases, a radial force (F z ) and a bending moment (M y ) are applied. However, in the first load case, no axial force (F x ) is applied, whereas an axial force is applied in the second load case. Science and Research 16 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0003 ing, and a cover with a pocket for exchanging the segments (Figure 2). The inner ring is mounted on the shaft. The bearing segments are mounted on the inner ring and therefore rotating. This bearing concept enables individual bearing segments to be replaced without disassembling the drivetrain. The width of the scaled specimen is w = 124 mm, and the sphere that defines the contour of the segments has a diameter of d = 340 mm. Various coatings are applicable to the bearing segments. The tests presented in this paper were carried out using segments with an aluminium-tin coating. All the experiments presented in this work are executed with the bearing in a state after the run-in. The bearing is lubricated with ISO VG 320 oil by a sump lubrication system at atmospheric pressure. To control the temperature of the oil, the oil sump is connected to an external oil circuit, which allows the oil to be cooled and heated. Load cases Low sliding speeds result in low pressure buildup in hydrodynamic plain bearings, leading to solid body con- Figure 3: Loads of a start-stop-cycle Figure 2: Structure of the spherical plain bearing [4] The sliding speed is set to 0.0890 m/ s at the start of each idling test and then gradually reduced to 0.0018 m/ s and held constant for several seconds at each speed step. 3 Results & Discussion First, the start-up after different drainage times is discussed. During drainage, the bearing is loaded at standstill, so that the lubricant film is minimized and solid contact occurs. The system starts after drainage times of between five minutes and 17 hours. Apart from the variation of the drainage time, the loads are identical to the loads in Figure 3. The progression of the measured torque over time after startup is shown for different drain age durations in Figure 5. The breakaway torque is between 3.88 kNm for five minutes drainage time and 4.77 kNm for 17 hours drainage time. The breakaway torque does not increase strictly with drainage time, but a trend can be recognized. Thus, the breakaway torque correlates with the drainage time. Schröder et al. also investigated a concept for plain main bearings and found good breakaway behavior [6]. In comparison, a similar breakaway torque can be determined for the Miba SJB™ under higher loads. When only the maximum torque is considered, the torque peaks in Figure 6 can be plotted over the drainage time. The data from a second test run with the same drainage times each is added here. The breakaway torque is between 3.8 kNm and 4.9 kNm for all tests. However, there is a noticeable correlation between drainage time and breakaway torque. In general, convergence of the breakaway torque can be assumed over longer drainage times. This is an advantageous Science and Research 17 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0003 Figure 5: Torque progression for start-up after different drainage-times Figure 4: Idling loads oil viscosity. One possible reason for this is a thinner lubrication gap and a larger load-bearing sliding surface at lower viscosities, which results in higher friction losses. Since the focus is on breakaway behavior at this point, this topic will not be discussed further. A radial plain bearing solely under vertical load will come to rest at the 6 o’clock position of its housing, with solid body contact between shaft and housing (Figure 8a). Radial plain bearings initiate their rotary motion with a rolling motion and the center of the shaft shifts horizontally by a distance s (Figure 8b) [9]. Only afterwards, the shaft begins to slide along the housing surface. The sliding motion initially occurs with solid contact, until the hydrodynamic pressure is sufficient to completely separate the sliding surfaces with lubricant (Figure 8c). The bearing under investigation mainly bears loads in the radial direction. To investigate whether a starting be- Science and Research 18 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0003 behavior of the bearing, as long downtimes do not lead to an extreme increase of the breakaway torque. The start-up behavior was also investigated at different oil temperatures to estimate the effect of oil viscosity on the breakaway torque. Again, the load case from Figure 3 is used. The bearing is preheated with oil before the test to ensure a constant temperature of the test specimen. Figure 7 shows the torque curve for the measured oil sump temperatures of 21, 28, 34 and 39 °C under otherwise identical conditions. The measured breakaway torque varies by a maximum of 0.13 kNm in the temperature range investigated. Therefore, the influence of temperature on breakaway torque can be assumed to be low. It is noticeable that the torque loss at higher speeds (on the right in Figure 7) increases with increasing temperature and thus decreasing Figure 7: Torque during start-up at different oil temperatures Figure 6: Breakaway torque after different Drainage-times havior similar to Figure 8 can also be observed on the spherical bearing, the displacement (s) of the shaft during system start-up is considered below (Figure 9). To avoid statistical effects due to the angular position of the bearing at the start-up, 30 consecutive system start-ups under the same loads are considered. Figure 9 shows the two tests in which the maximum and minimum displacements were measured. It plots the displacement s of the shaft over time. The dashed line marks the time at which the peak breakaway torque is reached. The applied torque initiates a horizontal movement of the shaft, starting at t = 0 s until the moment of breakaway at t = 2.5 … 2.7 s. Horizontal shaft movement during start up ranges between 5.4 μm and 8 μm. The displacement of the shaft before sliding begins between the outer ring and segment indicates a rolling motion similar to Figure 8. The shaft initially rolls along the outer ring, as illustrated in Figure 8. Sliding between the segments and the outer ring begins at the time of breakaway. Sliding is thus initiated in a position of lubricated contact, rather than in the drainage position, as the shaft first rolls out of the drainage position. This leads to a reduced breakaway torque due to less solid contact. This behavior was observed for all 30 start-ups. In the following, the load behavior during idling operation is discussed. The frictional torque is measured during idling operation. It must be noted that, in addition to the frictional torque of the test specimen, the torque transducer also measures torque losses of the bearing of the load triangle (red part in Figure 1) and the support bearing. To estimate the influence of the two rolling element bearings, tests without non-torque loads (NTL) are carried out, varying the speed in different stages. It is also assumed that the load-dependent loss in the plain bearing is significantly higher than in the two rolling element bearings. The maximum torque loss resulting from the three bearings (spherical plain bearing and two rolling element bearings) while no-load tests is less than 80 Nm. Since the measured torque under load in the low-speed range exceeds 1.2 kNm, the loss due to the two rolling element bearings is considered negligible. Figure 10 shows the relationship between the frictional torque and sliding speed qualitatively for both loadcases from Figure 4. The minimum of both curves lies in the respective highlighted area. The curve is similar to the typical Stribeck curve for plain bearings. In the second load case (idling 2), the transition speed is achieved at a lower sliding speed than in the first load Science and Research 19 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0003 Figure 9: Maximum and minimum shaft-displacement of 30 start-ups a b c s Figure 8: Rolling motion of a radial plain bearing according to [10] This paper examines the behavior of a segmented and spherical plain bearing during start-up and idling with scaled down tests. The tests were carried out on a test rig applying scaled WT forces and bending moments. The results show that the breakaway torque during start-up and demonstrate the impact of drainage times on breakaway torque. Systematic testing revealed a convergence of the breakaway torque after longer drainage times of up to 17 hours indicating that there is a limit, where the complete oil film is squeezed out and only solid body contact remains. The influence of oil temperature only has a minor influence on breakaway torque within the tested temperature range. Additionally, the bearing’s behavior during system start-up includes a rolling motion before breaking away. Rather than experiencing solid contact immediately upon start-up, there is a slight movement of the shaft along the circumferemce of the outer ring before starting to turn. Idling tests were used to study the bearing’s speed-dependent frictional torque. A transition sliding speed of between 0.018 m/ s and 0.052 m/ s was determined. It can Science and Research 20 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0003 case. The transition speed for the idling 1 load case is between 0.035 m/ s and 0.052 m/ s, and the transition speed for the idling 2 load case is between 0.018 m/ s and 0.035 m/ s. The transition speed is lower than that of comparable concepts [6, 11]. Additionally, the frictional torque for the second load case is lower than for the first at low sliding speeds. This is due to the asymmetrical bearing geometry, as seen in Figure 2 on the right. Under a slight axial load, the load-bearing area of the sliding surface increases as the load is pointing more towards the center of the sliding surface (Figure 11 (b)). Consequently, hydrodynamic pressure builds up at lower sliding speeds. This behavior is beneficial for WTs because high thrust forces can act on the rotor in the field, which results in an axial loading of the main bearing. The friction loss differs only minimally at higher sliding speeds. 4 Conclusion and Outlook Plain bearings as main bearings enable a reduction of LCOE of future WTs due to lower maintenance costs. Figure 10: Relationship between the frictional torque and sliding speed for idling loadcases Figure 11: Resulting force on the sliding surface for two different load cases thus be concluded that no solid body contact occurs even at comparatively low sliding speeds, enabling the WT to start up at low wind speeds. This suggests that the bearing enables the WT to start up at low wind speeds. Next, tests will be conducted with a full-size bearing for a 4.x MW WT proving the transferability of the scaled test results to larger designs. As with the small test specimen, idling and start-stop tests will be carried out. References [1] Bundesministerium für Wirtschaft und Klimaschutz: Erneuerbare Energien in Zahlen. Nationale und internationale Entwicklung im Jahr 2023 (2023). URL https: / / www.bmwk.de/ Redaktion/ DE/ Publikationen/ Energie/ 20241129-erneuerbare-energien-in-zahlen- 2023.pdf? __blob=publicationFile&v=12 - Überprüfungsdatum 2025-05-05 [2] Hart, Edward; Turnbull, Alan; Feuchtwang, Julian; Mcmillan, David; Golysheva, Evgenia; Elliott, Robin: Wind turbine main-bearing loading and wind field characteristics. In: Wind Energy 22 (2019), Nr. 11, S. 1534- 1547 [3] Dalgic, Yalcin; Lazakis, Iraklis.; Turan, Osman; Judah, Sol: Investigation of optimum jack-up vessel chartering strategy for offshore wind farm O&M activities. In: Ocean Engineering (2014) - Überprüfungsdatum 2025-05-05 [4] Hölzl, Johannes; Laubichler, Patrick; Mirza, Michael; Reisenberger, Johannes; Zeh, Christopher: Miba’s Segmented Journal Bearing (SJB™) - an innovative solution for wind turbine main bearings. In: Conference proceedings CWD, 2025 [5] Song, Chaosheng; Zhang, Ronghua; Zhou, Ye; Tan, Jianjun; Zhu, Caichao: Transient analysis of non-uniform tilting-pad journal bearing for wind turbine main shaft under dynamical loading. In: Mechanism and Machine Theory 203 (2024), S. 105792. URL https: / / www.sciencedirect.com/ science/ article/ pii/ s00941 14x24002192 [6] Schröder, Tim; Jacobs, Georg; Rolink, Amadeus; Bosse, Dennis: “FlexPad” - Innovative conical sliding bearing for the main shaft of wind turbines. In: Journal of Physics: Conference Series 1222 (2019), Nr. 1, S. 12026. URL https: / / iopscience.iop.org/ article/ 10.1088/ 1742-6596/ 1222/ 1/ 012026/ meta [7] Center for Wind Power Drives: NextMBU | Center for Wind Power Drives | RWTH Aachen University. URL https: / / www.cwd.rwth-aachen.de/ cms/ cwd/ forschung/ projekte/ ~vdjik/ nextmbu/ . - Aktualisierungsdatum: 2025-07-14 - Überprüfungsdatum 2025-07-14 [8] Poll, Gerhard; Deters, Ludger: Lagerungen, Gleitlager, Wälzlager. In: Sauer, Bernd (Hrsg.): Grundlagen von Maschinenelementen für Antriebsaufgaben. 8. Auflage. Berlin: Springer Vieweg, 2018 (Springer-Lehrbuch, 2), S. 69- 193 [9] König, Florian: Prognose des Verschleißverhaltens ölgeschmierter Gleitlager im Mischreibungsbetrieb. Dissertation. 2020 [10] Childs, T. H. C.: Mechanical Design: Theory and Applications. 3 rd ed. San Diego: Elsevier Science & Technology, 2021 [11] Hameyer, Kay; Jacobs, Georg; Monti, Antonello; Schröder, Wolfgang; Hirt, Andreas: Conference for Wind Power Drives, CWD 2019: Aachen 12 th - 13 th of March: conference proceedings, Aachen, 12 th - 13 th of March 2019. 1. Auflage. Norderstedt: Books on Demand, 2019 Science and Research 21 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0003 Science and Research 22 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 1 Introduction In modern technical science, the radial shaft seal (RSS) has been developed to match the industrial needs in various aspects. To verify and extend the potential lifetime limit, the function testing environment for RSS has been as well extended in various standards. This work presents a new testing environment for RSS, specifically focusing on the potential application of 3D printed counterpart following from [1]. After different 3D printing processes, different initial materials, and various postprocessing methods, the 3D printed counterparts have been investigated with a reference oil and one specific RSS with various operation conditions. The results show partially comparable results with the standard counterpart, which leads to potential applications of 3D printed counterparts in research field. 2 Test setup and test material In this section, the test procedure, more specifically the applied test setup and test material, will be introduced one after another. Similar like the previous works, the specially designed RSS test rig at the Institute for Machine Elements, Design and Manufacturing has been implemented for the experimental tribological investigations, which provides comparable operation conditions while using simplified minimal lubrication principle [2] [1]. After that, in section 2.2 the test materials are intro- Tribological Investigations of Radial Shaft Seal with Focus on 3D Printed Counter Surface Yongzhen Lin, Matthias Kröger, Jenny Köckritz, Deeksha Lohani, Ringo Nepp* submitted: 26.09.2025 accepted: 19.03.2026 (peer review) Presented at GfT Conference 2025 Tribological performance of radial shaft seals (RSS) has been evaluated among experimental investigations with different metal counter surfaces in the recent decades. The counter surface has been identified as one of the main drivers determining the tribological performance of RSS. In general, the metal counterpart needs to be manufactured specifically to achieve a specific surface structure as well as an optimal surface roughness. The appropriate application of RSS as a dynamic seal element in different types is depending on different boundary conditions. In order to extend the knowledge on the boundary conditions specifically the potential application of RSS on 3D printed counterparts, tribological investigations have been performed and evaluated in this work. More specifically,3D printed counterparts with different materials, various 3D printing processes with or without postprocessing have been examined. The potential operation limits such as temperature and rotation speed have been searched. Correlations between the counter surface properties and the tribological performance have been observed and evaluated from different aspects and in different ways. The results of this work are determined experimentally, particularly focusing on the methodical approach on 3D printed counterparts. A specifically established RSS test rig at the Institute for Machine Elements, Design and Manufacturing is available for the experimental tribological investigations. The aim of this work is to set up a basis for the future researches and potential applications regarding optimization operation conditions of RSS for wear reduction and service life extension. As well as gaining new insights for the possibility of 3D printed counterparts in tribological investigation of RSS. Keywords Radial shaft seal, counterpart, 3D manufacturing, new material, tribological investigation, surface topography analysis Abstract * Yongzhen Lin, M. Sc. (corresponding author) Orcid-ID: https: / / orcid.org/ 0000-0002-0700-0914 Prof. Dr.-Ing. Matthias Kröger Orcid-ID: https: / / orcid.org/ 0000-0002-4132-8323 Dipl.-Ing. Jenny Köckritz Orcid-ID: https: / / orcid.org/ 0009-0000-6042-5367 Dr. Deeksha Lohani Orcid-ID: https: / / orcid.org/ 0009-0003-8455-8473 Dr.-Ing. Ringo Nepp Orcid-ID: https: / / orcid.org/ 0009-0001-7837-258X TU Bergakademie Freiberg Institute for Machine Elements Design and Manufacturing Agricolastraße 1, D-09599 Freiberg duced, especially the applied initial materials for 3D printing processes. Analog from one previous work [1], the 3D printed parts are inserted into tribological investigation to get the tribological properties such as friction torque and temperature. 2.1 Test setup for tribological measurement A specially designed RSS test rig at the Institute for Machine Elements, Design and Manufacturing serves for tribological performance measurements of RSS under minimal lubrication concept. This test rig and the relevant components have been described in Figure 1. In this study the shaft part has been replaced with different 3D printed counterparts. This test rig has been operated for different tribological investigations at the institute to extend the knowledge of tribological behavior of RSS under various influences such as different lubricants, different environment temperature [2]. 2.2 Test material The applied test materials including the RSS, the lubricants as well as the shaft counterparts are introduced in detail in this section. In order to keep the variations of the whole tribological system as simplified, the variations in this study are placed on the shaft counterparts using different materials and potentially the different circumferential velocities. 2.2.1 Radial shaft seal and lubricants In this work, the focus is placed on the potential influence of different counter surfaces. The key parameters of the applied materials are listed in Table 1. In order to keep the focus of this study on potential applications in research of counterparts using different initial materials, the other important components such as RSS and lubricant have been selected as standard ones after the industrial standard [3] [4]. Among all the materials of counterpart, the C45R is selected as a standard reference counterpart, while stainless steel 316L is laser powder [5], Thermoplastic Polyurethane (TPU) is a plastic filament [6] and Copper* filament (Copper*) is indicated as one specific copper filament with over 80 % copper pigment and 20 % Poly Lactic Acid (PLA) [7]. These are initial 3D printing materials with different composites. The states of counterparts are categorized as standard, As-build (AB), Mechanical processing (MP) and Plasma electrolytic polishing (PeP). The process Plasma Electrolytic Polishing (PeP) is an advanced, non-conventional electrochemical surface finishing technique characterized by applying high voltage between the workpiece and an electrolyte, inducing localized micro-discharges at the interface. These micro-discharges facilitate selective anodic dissolution of surface asperities, resulting in substantial surface smoothening while maintaining high dimensional accuracy. The process typically achieves material removal on the order of micrometers, ensuring minimal impact on part geometry. PeP can produce ultra-smooth, mirrorlike finishes across a wide range of metallic materials, including stainless steels, titanium alloys, and aluminumbased substrates. Its non-contact, electrochemical mechanism allows efficient polishing of complex geometries, intricate internal channels, and hard-to-reach features that are challenging for conventional mechanical polishing methods. From an environmental perspective, PeP is a sustainable surface finishing approach, as it employs non-toxic electrolyte soluti- Science and Research 23 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 Figure 1: Applied test rig with necessary components. "#$%#&'&()! *+(',-+.)! +&/ ! 0,#%',(-')! 122! 341! 456*7! 78! 9! : ; ! 9! <=! >: ! 9! <: ! 9! <! ? @A,-B+&(! CD57! E! >F5! G2H! DI! JKK! LG2H! 7>>8M! N'&)-(O! 88>PJQRST$ ! " LJ: QU"M "" ! D-)B#)-(O! VKPK; Q$$ # T)! L>KQU"M! ! ! ! ! ! ! ! ! ! ! ! ! ! ! ! JKP>JQ$$ # T)! LJKKQU"M! "#@&(',%+,(! ">: 1! 2(+&/ +,/ =! S,#@&/ ! 7JW? ! 54=! *0=! 0'0! X06! 54! "#%%',Y! 54=! *0! Table 1: Key parameters of the applied materials. FFF process the applied materials are TPU and Copper*. The production processes for different materials also with different 3D printers differ from each other. The important printing parameters from different 3D printing machines are listed in Table 3. The different printing methods influence to a certain extent the material properties from the initial state to the product part. In Figure 2, surface structure differences on the As- Built (AB) parts from different materials can be observed significantly. From 316L and SLM procedure, the AB part has clearly layered structure probably due to the infill scan strategy with 67°. This structure is not clearly visible on the AB TPU part and FFF procedure. Meanwhile the AB Copper* part shows a completely different surface structure, on which there are clearly powdery particles visible. After production, some of the AB metal samples have been processed with two different post-processing methods, which are MP and PeP. Since it’s difficult to apply post-processing on TPU part, the TPU samples are applied in tribological investigation in AB state. Other than that, the 316L and Copper* parts are post-processed. Some 316L samples remain also not post-processed to compare the surface structures under microscopic analyse. Before PeP, the samples were Particle blasted at 9000 rpm for 30 min to ensure uniform surface pre-conditioning. PeP was subsequently carried out at an applied voltage of 350 V and an electrolyte temperature of 80 °C. The treatment was performed in four cycles of 10 min each, with the sample rotated by 90° after every cycle. This procedure resulted in a total polishing duration of 40 min, ensuring complete 360° surface coverage. Science and Research 24 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 ons and generates minimal chemical waste or gaseous emissions. Furthermore, the absence of mechanical contact eliminates tool wear, contributing to process stability and repeatability. Despite these advantages, PeP presents challenges in achieving precise local control of material removal. The stochastic nature of plasma formation and micro-discharge distribution can result in non-uniform polishing at localized regions, which may limit its application for components requiring extremely tight tolerances in specific areas. Continued research is therefore focused on optimizing electrolyte composition, voltage parameters, and process dynamics to enhance local control while retaining the inherent benefits of the technique. To have a short insight on the mechanical characters of 3D printing materials, the physical properties of the standard part and the other initial materials are summarized in Table 2. The Youngs’ modulus of 316L has been measured at the institute according to [8] methodically. The Youngs’ modulus of other materials are from industrial data sheets [6] [7] [9]. As it could be observed that the 316L powder has a much higher Youngs’ modulus than Copper* and TPU materials. Meanwhile, 316L has also a relative higher thermal conductivity in comparison with TPU filament. The thermal conductivity of the Copper* filament remains unknown. 2.2.2 Shaft counter faces Using additive manufacturing methods, multiple components can be produced within hours. The applied methods in this work are Selective Laser Melting (SLM) metal and Fused Filament Fabrication (FFF) plastics. For SLM process the applied material is 316L, while for C45R 316L Copper* TPU Youngs’ modulus [MPa] 210*10 3 180*10 3 4210 40 Thermal conductivity [W/ K*m ] 42.6 15 - 0.19 Table 2: Material properties of the applied counterparts. raw material 316L TPU Copper * 3D printing Machine Farsoon © FS273M Creality Ender 3 S1Pro Bambu Lab © A1 Proceeding SLM FFF FFF Nozzle temp 220 °C 200 °C Print plate temp 100 °C 60 °C 60 °C Layer thinkness 0.041 mm 0.25 mm 0.08 mm First layer thinkness - 0.3 mm 0.1 mm Infill density 100% 100% 100% Infill pattern Scan strategy_X, 67 ° Straight linear Straight linear First layer velocity - 30 mm/ s 50 mm/ s Infill laser velocity 900 mm/ s 30 mm/ s 105 mm/ s Table 3: 3D production processes for different initial materials. ab © A1 near s ab © A1 near s Youngs’ modulus [MP Thermal conductivity [W s’ modulus al conduct In Figure 3, the post-processed 316L parts are shown in comparison with the AB part. As it is directly observed that the MP and PeP processes change the surface structure significantly, while the surface structures changing after MP and PeP differ from each other as well. These optical differences raise the question if the surface roughness ranges and the material properties are sufficient to be applied in tribological investigation of RSS. 2.3 Measuring procedure The mean measuring procedures for this paper are first the surface roughness measurement after the industrial standard [10] [11] [12] and second the tribological investigation. From the topography surface measurement, not only the roughness parameters have been summarized, but also the topographies show the difference in surface structures from different materials and different post-processing. 2.3.1 Surface roughness measurement Before the parts are inserted into tribological investigations, the surface topography and surface roughness properties have been measured and evaluated. In Figure 4 and Figure 5 are the topography results of the parts. From the topography figures in Figure 4, clear difference among AB part and post-processed parts can be observed. On AB part there are visible micro particles. PeP process changed the surface character completely. On MP part there are visible polishing traces. In Figure 5 are the topography results of three different materials. Here the different surface structures are clearly visible. The Copper*-MP part shows slight different surface structure than 316L- MP part, mostly due to the initial material and the printing process. Other than that, the TPU part with the layer structure is clearly visible, which is strongly influenced by the filling ratio, the melting temperature as well as the filling velocity. In addition to this, the surface roughness has been measured to reach the surface tolerance according to [12]. As it’s shown in Figure 6, the surface roughness parameters are shown for all the measurements including the AB samples. For all parts, six parameters Science and Research 25 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 AB MP PeP Figure 4: 3D topography figures from the 316L metal samples with and without post-processing. 316L TPU Copper* Figure 2: Surface structure on AB parts from different materials and procedures. Figure 3: Optical observation on AB part and post-processed 316L parts. have been removed. For further analysis in this study, it is focused on R i values. Since the surface structure and the profile valley depth could influence the lubricating film and further the tribological performance, it’s interesting to observe from the diagram in Figure 7 that there is extremely large difference between the TPU part and the metal parts. Other than that, among the metal parts, the Copper* part shows as well higher R z and R v values, which affect the tribological performance significantly. Beside the topography analyses, the concentricity tolerance of the parts has been verified after the industrial standard [15] [12]. This has been investigated with Hexagon Quindos © along the cylinder circumference. 50 sampling points have been taken from each part. The results are listed in Table 4 as follows. The actual diameter D A and the actual roundness R o of the test parts have been measured by 50 measuring points along the circumference. The results are listed in Science and Research 26 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 are shown in diagram, which are specific: Arithmetical mean height (R a ), Maximum height of profile (R z ), Maximum profile valley depth (R v ), Root mean square deviation (R q ), Medium arithmetic height (S a ) and Medium square height (S q ) [13]. All these values describe the surface character from different aspects. In general, R i values describe the measured values along with the sampling length, while S i values stand for the measured values within a defined area range [14]. It’s clearly to observe that the AB parts from three different materials demonstrate higher surface roughness values, especially the R z and R v values. At this point, it needs to be highlighted that the post-processing helps to improve the surface quality significantly. Especially the parts D45-316-PeP have a R a 0.65 μm, R z 3,06 μm while the D45-316L-MP parts have a R a 0.97 μm R z 5.94 μm. Figure 7 shows the surface roughness measurement results of the parts that have been investigated in tribological experiments afterwards. The results of AB parts Figure 5: Topography results of three different initial materials. Figure 6: Surface roughness measurement results from all the parts with and without post-processing. #$$ & #$$ & )& &#) " '' -! )& &#) " '' ! ')& ! "( H * % % Table 4. The industrial Standard [12] and [15] require the shaft diameter tolerance ΔD t should not be exceeding 0.16 mm and the roundness tolerance R ot should not be exceeding 0.02 mm. It can be observed that the D45-316L-MP and D45- 316L-PeP parts have the allowable tolerance on diameter and roundness, while the D38-Copper*-MP part exceeds slightly the roundness tolerance, but has a good fit on diameter tolerance. Nevertheless, the D38-TPU part has exceeded the tolerance on diameter and the roundness tolerance a lot. 2.3.2 Tribological investigation The tribological performance has been carried out through experimental investigations on the RSS test rig. To explore the potential applications for function tests, a parameter study has been carried out to select the maximum boundary. In Table 5, the applied circumferential velocities have been generated from 1 m/ s to 5 m/ s in various steps. The study [16] has shown that various velocity steps have a significant influence on tribological performance of RSS with specific lubricants. Generally, the tests have been running at a constant circumferential velocity for 1 or 2 hours to reach the stationary states. In addition, multiple repeated tests have been carried out to validate. In Table 5 are the applied circumferential velocities from 1 m/ s to 5 m/ s labeled from E11 to E19. The investigations have been carried out from lower circumferential velocity. The velocity for each test has Science and Research 27 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 Figure 7: Surface roughness measurement results from all parts applied afterwards in tribological investigations. ! "" # &# #! & $$ $&# % ( ' D A [mm] ΔD [mm] ΔD t [mm] R o [mm] R ot [mm] D38-TPU 37.583 0.471 0.16 0.064 0.02 D38-Copper * -MP 37.843 0.157 0.16 0.027 0.02 D38-316L-MP 38.023 0.023 0.16 0.008 0.02 D45-316L-MP 44.942 0.058 0.16 0.01 0.02 D45-316L-PeP 44.901 0.099 0.16 0.02 0.02 Velocity [m/ s] D38, Standard D38-TPU- AB D38- Copper * - MP D38-316L- MP D45-316L- MP D45-316L- PeP E11 1 2h E12 1.2 2h 1h / 2h E13 1.4 1h / 2h 1h / 2h 1h / 2h 2h E14 2 broken 2h 1h / 2h 2h E15 2.1 2h 1h / 2h 2h E16 2.4 broken 1h / 2h 2h 2h E17 2.8 1h / 2h unstable 2h E18 3.5 1h / 2h unstable E19 5 2h unstable Table 4: Actual diameter and actual roundness measurement. Table 5: Applied circumferential velocity profiles and the examined profiles. in comparison with the part D38-Copper*-MP, while the part D38-316L-MP has determined higher torque and temperature values. In addition, the part D38-316L-MP shows that torque and temperature reach slowly stationary state after 30 min, with a temperature about 103 °C near the contact area. Meanwhile the results of part D38- Copper*-MP show a rapid heat up and it reaches an almost stationary state in temperature immediately after 3 min of the velocity application, with a temperature of about 50 °C. Afterwords, the temperature increases slowly till the end of measurement. The torque development shows similarity from both test conditions, yet the absolute values differ from each other. Characteristic results have been marked with three circles in the left diagram in Figure 8, the mean values of the maximal torque at t 0 , the minimal torque at t end and the temperature at t end . These results from each measurement are chosen for the following evaluation. In the very beginning, the maximal torque at t 0 and minimal torque at t end have been presented in Figure 9. These values have been generated from part D38-TPU, D38- Copper*-MP, D38-316L-MP and D45-316L-MP. As it Science and Research 28 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 either been repeated or has been added up to next step. The adding step is stopped when the seal or the counterpart fails. After several repeat tests, the tribological performance could be evaluated based on the measurement results. TPU and Copper* parts are broken after the defined maximal velocities, while the 316L parts remines in good states. In addition, the torque and temperature developments with 316L parts are not stationary with the defined maximal velocities. 3 Results and discussion The measurement results are discussed in this section considering different aspects. First, examples of a whole measurement after 1 hour of two specific tests are shown in Figure 8. Analog to [17] the local tribology of the RSS has been evaluated. On the left, the diagram shows a measurement result from part D38-316L-MP with a constant rotation velocity 3.5 m/ s, while on the right side, the diagram shows a measurement result from D38- Copper*-MP with a constant rotation velocity 1.2 m/ s. From both diagrams, it can be observed that the part D38-316L-MP can stand higher circumferential velocity Figure 8: Torque and temperature measurement of one specific test. Left: D38-316L-MP at 3.5 m/ s, right: D38-Copper*-MP at 1.2 m/ s. Figure 9: Maximal and minimal torque of all investigations from different parts measured in 1 hour. "%$( " ') & ! '"%$( ' "## % can be observed from the diagram, the TPU and Copper* parts can only endure less than 2 m/ s. Although the parts D38 and D45 have been post-processing with MP, they have different circumferential velocity tolerance and different torque performance, mostly due to the difference in surface structure, which could be summarized from the topography measurement in Figure 7. In Figure 10 the temperature at t end from all the tests has been presented. These points are generated either from different circumferential velocities or from repeat tests. As it can be observed that from one material, the temperature range is relatively large, especially on the part 316L, the temperature range is over 30 °C on velocity difference of 1 m/ s. On the other side, tests with TPU parts have shown higher temperature than Copper* having the same friction torque (see Figure 9). This shows different thermal properties of both these materials. After the evaluation of measurement results of 1 hour, it could be deduced that the testing time is too short to observe any potential long-term changes. Therefore, several tests in 2 hours have been carried out. The testing principle remines to extend the potential maximal velocity, until the counterpart or the seal fails. The evaluation of the tribological performance is based on the minimal torque at t end and temperature at t end . In this batch of tests, the D38-Standard counterpart has been tested to have a reference for the other measured results. All the test results are presented in Figure 11. As it can be observed from the diagram, only the standard counterpart can stand higher velocity of 5 m/ s. The D38-316L-MP parts can stand relatively high velocity till 3.5 m/ s. Between the D45 parts with two different post-processing methods, the PeP part can stand relatively higher velocity till nearly 3 m/ s, while the MP parts stand around 2 m/ s. Last but not less, the TPU part can stand lower than 2 m/ s and the Copper* parts can stand slightly higher than 2 m/ s. In comparison with TPU, the Copper* parts can stand higher velocities, and cause lower temperature. The 316L parts both with D38 and D45 cause higher torque and stranger temperature increase with lower velocities then the reference part C45R. In comparison with TPU and Copper* parts, which cause lower torque and lower thermos effects at lower velocities. All measured temperature results are shown in Figure 12. From this diagram, it can be observed that the TPU, Copper*, D45-316L- MP and D45-316L-PeP parts with lower velocity have caused linear proportionally lower temperature changes. Meanwhile the D45-316L-PeP as well as the D38-316L-MP with higher velocity have caused higher temperature changes. The results of tests with D38-Standard show lower temperature range lower than 100 °C at higher velocities, which could be explained with higher thermal conductivity. Other than that, the friction coefficient μ is calculated with formula μ = 2 * M min / (D * F R ), Science and Research 29 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 Figure 11: Maximal and minimal torque of all investigations from different parts measured in 2 hours. ! $'! ( *+(! - ! #& "*, $ ) ! $' * * ! % &''! ( "%! ( &''! ( "%! ( "%! ( $'&* " ! $ )+ " ( # )$'&* ) $%% ' ) # ' Figure 10: Temperature scattering from adding up circumferential velocities measured in 1 hour. different heating processes with different materials could be observed with more details. Other than that, the rotation speed profile could also to be extended such as inconstant, alternately changing profile to extend different tribological performance testing among different testing environments. References [1] Y. Lin, R. Nepp, J. Gerschler and M. Kröger, “Radial shaft seal with different relative velocities and on a 3D printed counter surface,” Industrial Lubrication and Tribology, pp. 758-764, 75(7) 2023. [2] C. Berndt, Tribologie von Radial-Wellendichtungen, Dissertation, TU Freiberg, 2019. [3] Weber Reference Oils, FVA3A Referenzöl Datenblatt, 2018. [4] Deutsches Institut für Normung e.V. (Hrsg.), DIN 3760: 1996-09: Radial-Wellendichtringe, Berlin: Beuth, 1996. [5] eos, “EOS StainlessSteel 316L-4441,“. Available: https: / / store.eos.info/ de/ products/ eos-stainlesssteel-316l. [accessed on Aug. 26. 2025]. Science and Research 30 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 in which the M min refers minimal torque at t end , D refers the nominal diameter and F R refers radial force which has been measured on new RSS before utilization. The results are shown in Figure 13. As it can be observed from Figure 13, TPU and Copper* parts have caused less friction in the contact area at lower velocities. On the other hand, the 316L parts with MP and PeP post-processing have contributed to higher friction. The standard part shows medium friction coefficient. 4 Summary and conclusion In this work, tribological investigations have been carried out with the focus on different counterparts, especially with 3D printed parts from different materials, different printing processes and different post-processing methods. Before the tribological investigation, surface roughness and other essential measurements have been executed to get more knowledge about the surface structure as well as other mechanical properties. The tribological performance has been tested with different constant circumferential velocity profiles within 1 hour or 2 hours. The constant circumferential velocity varies from 1 m/ s to 5 m/ s. After the experimental investigations, the measured torque and temperature values from all the counterparts have been evaluated in various aspects. The measurement results show significant tendencies of the relevance of the torque temperature depending on the velocity with counterparts of different materials. This study provides a new effective trend at the institute internally by applying additive manufactured parts from new materials as well as new producing processes also in tribological applications. This would help a lot, especially in the first prototype testing phase of products whether for new RSS types, new lubricants, new test procedures or new counterparts. In the future, the thermal properties of different 3D printing initial materials need to be further measured and evaluated. These properties have significant influences on the tribological investigations of RSS. Meanwhile, ! $'! ( *+(! - ! #& "*, $ ) ! $' * &''! ( * % ( ! "%! ( "%! ( ! "%! ( Figure 12: Temperature scattering from adding up circumferential velocities measured in 2 hours. (" *"&% & ! #& "*+ $ ) (" *"&% & ! * &'' ( * % ( Figure 13: Friction coeffient of all investigations from different parts in 2 hours. [6] extrudr, “FLEX Hard CF,”. Available: https: / / www.extrudr.com/ de/ de/ products/ flex-hard-cf/ ? variant=UHJvZHVjdFZhcmlhbnQ6MzYyOQ%3D%3D. [accessed on Aug. 26. 2025]. [7] eins3d, “Kupfer Filament,”. Available: https: / / eins3d.de/ products/ kupfer-filament-2-85mm-500g. [accessed on Aug. 26. 2025]. [8] C.-H. Yu, A. Leicht, R. L. Peng and J. Moverare, “Low cycle fatigue of additively manufactured thin-walled stainless steel 316L,” Materials Science and Engineering: A, p. V. 821, 2021. [9] Deutsche Edelstahlwerke, “Werkstoffdatenblatt C45R,”. Available: https: / / swiss steel-group.com/ content-media/ documents/ Data-Sheets/ Engineering-Steel/ 1.1191_1.1201_de.pdf. [accessed on Aug. 26. 2025]. [10] Deutsches Institut für Normung e.V. (Hrsg.), DIN ISO 25178-2: Geometrical product specifications (GPS) - surface texture: Areal, 2016. [11] W. Tietze, Handbuch Dichtungspraxis. 3.Aufl., Essen: Vulkan-Verl., 2003. [12] Deutsches Institut für Normung e.V. (Hrsg.), DIN3761: Radial-Wellendichtringe für Kraftfahrzeuge, 1984. [13 Keyence, “Introduction to roughness,” Available: https: / / www.keyence.com/ ss/ products/ microscope/ rough ness/ . [accessed on Aug. 20. 2025]. [14] M. Marxer, C. Bach and C. P. Keferstein, Fertigungsmesstechnik. Alles zu Messunsicherheit, konventioneller Messtechnik und Multisensorik. 10. Aufl., Wiesbaden: Springer Vieweg, 2021. [15] Deutsches Institut für Normung e.V. (Hrsg.), DIN EN ISO 1101: Geometrical product specifications (GPS) - Geometrical tolerancing., 2017. [16] M. Kröger, R. Teichert and Y. Lin, “Scaling of radial seals - Safely go into series production,” in JAHRBUCH Dichten. Kleben. Polymer., ISGATEC, 2024, pp. 80-90. [17] M. Kröger, Y. Lin, R. Nepp and S. Haupt, “Observation of local tribology of shaft seals,” in Arnold Tross Colloquium, Hamburg, 2024. Science and Research 31 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0004 the available design space. Consequently, adjusting fluid mechanical properties is generally not a promising strategy when relying on fossil-based diesel fuels. [7-9] This limitation is alleviated with the use of bio-hybrid fuels. The wide range of possible feedstocks, including biomass and by-products from the chemical industry, not only enables closed carbon cycles but also introduces substantial flexibility in fuel design. As a result, both combustion-related properties (e.g., cetane number) and fluid mechanical as well as thermal properties can be tailored over a broad range. In the context of drop-in fuels, improving pump efficiency through targeted fuel design is particularly attractive, as it does not require modifications to existing engine hardware. [10-13] Against this background, the present study investigates the potential to improve the efficiency of diesel injection pumps through the targeted blending of bio-hybrid fuels, with a focus on tailoring their fluid mechanical properties. To this end, a thermo-elasto-hydrodynamic (TEHD) model of the piston-bushing contact within a diesel injection pump is developed to quantify frictional and leakage losses. In addition, further performance indicators are evaluated to ensure that fuel blending does not adversely affect other system characteristics, such as the injected energy per unit time or the injection process itself. Science and Research 32 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0005 Introduction A considerable share of the power in diesel engines is required to drive the fuel injection pump, reaching several kilowatts even in mid-range vehicles. At the same time, substantial volumetric and hydromechanical losses occur within the pump. These losses become increasingly significant at elevated injection pressures and can result in comparatively low overall efficiencies, in some cases falling below 65 %. [1, 2] The use of bio-hybrid fuels may further exacerbate this issue, as their typically lower viscosity and reduced lubricity compared to conventional diesel fuels can lead to increased leakage and higher friction in tribological contacts within the pump. [2] Various approaches have therefore been proposed to improve the efficiency of injection systems. For instance, the studies by Heitzig and Rückert focus on the optimization of piston-bushing geometries. It has been demonstrated that gap compensation achieved by hollow piston designs can significantly improve volumetric efficiency, particularly at high pressures. Additional measures include piston contouring and the introduction of grooves to reduce leakage losses. Frictional losses, in turn, can be mitigated through the application of suitable coatings to tribological interfaces. [2-6] On the fluid side, friction and wear can be improved by the use of additives. However, enhancing overall efficiency through modifications of the fuel composition is of limited relevance for conventional diesel fuels, as their properties are tightly constrained by standards such as EN 590, ISO 4113, and ISO 12156-1. These standards impose narrow limits on key parameters including density, viscosity, and lubricity, thereby severely restricting Increasing efficiency of diesel injection pumps by targeted blending of bio-hybrid fuels Marius Hofmeister, Katharina Schmitz* submitted: 29.09.2025 accepted: 23.04.2026 (peer review) Presented at GfT Conference 2025 The efficiency of diesel injection pumps can be improved by adjusting the fluid mechanical properties of bio-hybrid fuels. Using a TEHD simulation model, this study investigates the impact of targeted fuel blending on friction and leakage losses, aiming to optimize fuel design for enhanced injection system performance Keywords TEHD, simulation model, Fluid mechanics, fuel design, enhanced injection system performance Abstract * Marius Hofmeister, M. Sc. Univ.-Prof. Dr.-Ing. Katharina Schmitz Institut für fluidtechnische Antriebe und Systeme (ifas) der RWTH Aachen Campus-Boulevard 30, 52074 Aachen Based on this simulation framework, different blending ratios of selected fuels are analyzed with respect to their influence on friction and leakage losses. The objective is to identify particularly sensitive parameters, such as pressure-viscosity behavior, that have a decisive impact on system efficiency. In a subsequent step, these parameters may be incorporated into the fuel design process to enable the development of optimized fuels that enhance both the efficiency and the reliability of diesel injection pumps. Injection system In Figure 1 the typical design of a common rail injection system is visible. The injection pump pressurizes the fuel and pumps it from the low-pressure line (blue) into the common rail, from where the fuel is injected into the combustion chamber via the injectors. Unlike common hydraulic machines, the main task of an injection pump is not to generate hydraulic power, but to provide a flow of chemical energy Q˙ heat that can be fed into the combustion process. This energy flow is equal to the product of pump volume flow V˙, the density of the fuel ρ l , and its calorific value H (eqn. 1). Losses due to internal leakage are considered by the volumetric efficiency η vol . To generate Q˙ heat , mechanical energy is required, which is equal to the product of the pressure difference Δp and the volume flow V˙. In this study, the ratio Q˙ heat and of P mech is defined as efficiency Φ and is used to describe how effectively mechanical energy is converted into a flux of chemical energy. eqn. 1 = ̇ ℎ ℎ = ̇∙ ∙ ∙ ∆ ∙ ̇∙ 1 ℎ = ∙ ℎ ∙ ∙ ∆ In addition to supplying chemical energy, injection pumps fulfill further secondary tasks. The high pressure in modern injection pumps is necessary to achieve a high spray velocity and, consequently, a high level of atomization, which has a positive impact on evaporation in the combustion chamber and subsequent combustion. High spray velocities also promote the mixing of fuel and air, thereby improving combustion efficiency. Finally, the hydraulic energy provided by the pump is partially converted into thermal energy in the injector, which also makes the heated fuel evaporate more easily. [14, 15] Consequently, a comparison of Φ alone is not sufficient for a meaningful assessment of pump efficiency, so that the secondary pump functions shall be considered, too. To describe the injection process with the listed phenomena quantitatively, extensive CFD models are necessary, such as those developed within the Fuel Science Center [15, 16]. In this study, atomization will initially be described qualitatively using dimensionless parameters such as the Weber and Reynolds numbers (We l , R e ) from eqn. 2 and eqn. 3. The Weber number We l is the ratio of inertial forces to surface forces and depends on the density ρ l and surface tension σ of the fuel. The relative fluid velocity u rel and nozzle diameter d are also included in the calculation. [14] eqn. 2 The Reynolds number Re represents the ratio of inertial forces to viscous forces. The inertial term corresponds to the product of relative velocity u rel , density ρ l , and nozzle diameter d. The viscous term corresponds to the dynamic viscosity η. [14] = inertial force suface force = 2 = 2 ∆ ∙ 2 ∙ Science and Research 33 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0005 Figure 1: Diesel injection system within this clearance can be accurately described by a thermo-elasto-hydrodynamic (TEHD) model. This model consists of a coupled solution of the Reynolds equation and the energy equation with variable fluid properties, while also accounting for mechanical and thermal interactions of the solid boundaries. Due to substantial heat generation and pronounced temperature gradients across the gap height, the Reynolds equation in the formulation of Dowson was applied, allowing viscosity variations in the direction normal to the gap. [17] In the present study, a similarly coupled model based on the Reynolds equation (eqn. 5) and the energy equation (eqn. 6) with variable fluid properties is employed. However, several simplifications are introduced, as the primary objective is to obtain an initial estimate of the influence of fuel blending on pump efficiency. In particular, the classical form of the Reynolds equation is used without incorporating Dowson-type integrals, and all forms of solid mechanical and thermal deformation are neglected. For the computation of the pressure field within the piston-bushing clearance, the Reynolds equation given in Eq. (5) is applied. [19] To solve the Reynolds and energy equations, a sufficient number of boundary conditions must be specified. The piston velocity is required for the shear flow. Pressure boundary conditions are required on both the high-pressure and low-pressure sides of the contact to calculate the pressure flow. A temperature condition on the high-pressure side of the contact is required to solve the temperature field. The boundary conditions T HP and p LP are specified at the beginning of the simulation, while the piston velocity and high pressure must first be calculated. The piston velocity is calculated based on the cam/ roller kinematics. To do this, the radius r C of the cam was measured at discrete points at intervals of 5°. A Fourier approximation with 10 terms was then performed to obtain a continuous relationship between the angle ωt and the cam radius r C (eqn. 7). o establish a function for the coordinates of the contact point (x c , y c ) between the roller and the cam, the system is converted into Cartesian coordinates. The point of ( ℎ ) ��� Pressure build−up = � ℎ 3 12 � ��������� Pressure flow − � ℎ ( 1 − 2 ) 2 � − 1 ℎ ������������������� Shear flow � + + � ����������������� Convection . − � 2 2 + 2 2 � ����������� = � + + � ��������������� Compression and expansion − �� � 2 + � � 2 � ������������� Science and Research 34 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0005 eqn. 3 The relative velocity u rel can be estimated by means of the nozzle equation and depends on pressure difference Δp, liquid density ρ l and flow coefficient α d (eqn. 4). Therefore, both dimensionless numbers can be represented as a function of the pressure difference Δp at the injector. eqn. 4 Equation eqn. 1 shows that the efficiency Φ depends on the fluid properties of the fuel, but also on the volumetric and hydromechanical losses (η vol , η hm ) of the injection pump. The losses occur in the tribological contacts indicated by the orange color in Figure 1. While hydromechanical losses mainly occur in the journal bearings of the shaft and the cam gear, the contact between the piston and the bushing is responsible for the volumetric losses. To evaluate the efficiency Φ of an injection system for different fuel combinations, the underlying loss mechanisms must first be quantified. In principle, this can be achieved using experimental methods, for example by conducting efficiency measurements based on established test procedures (Heitzig). However, given the large number of biohybrid fuel candidates investigated within the Fuel Science Center and the vast number of possible blending combinations, the experimental effort in terms of time and measurement complexity becomes impractical. [2] In addition to experimental approaches, efficiency can also be investigated using simulation-based methods [2, 17]. Due to their significantly lower time and resource requirements, a simulation-based approach is adopted in this study. Within this framework, the analysis is initially restricted to the piston-bushing contact, as it represents one of the dominant sources of frictional and leakage losses in the injection pump. The governing equations describing this contact, along with the relevant fuel properties, are introduced in the following sections. Fluid mechanics fundamentals When simulating the piston-bushing contact in a diesel injection pump, the extreme operating conditions impose specific requirements on the modeling approach. The wide ranges of pressure and temperature-spanning from ambient pressure up to 3000 bar and from approximately 40 °C to 200 °C - lead to significant variations in the thermophysical properties of the fuel, which can therefore not be assumed constant [18]. In addition, these high pressures and temperatures induce solid deformation, resulting in a non-uniform and variable gap height between piston and bushing. Fischer demonstrated, using a micro-gap representative of an unwrapped piston-bushing contact, that the flow = inertial force viscous force = = � 2 ∆ ∙ = � 2 ∆ ∙ eqn. 5 eqn. 6 contact between the roller (R) and the cam (C) can be found using the conditions (∂x/ ∂y) C = (∂x/ ∂y) R , and x c,m = 0 (Figure 2). eqn. 7 This allows a relationship between angle ωt and piston position y to be established (eqn. 8). By forming the time derivative, a relationship between piston velocity ẏ and angle ωt is obtained (eqn. 9). eqn. 8 eqn. 9 The pressure boundary condition on the high-pressure side results from the compression of the fuel in the piston chamber, which is modeled using the pressure build-up equation from eqn. 10. The bulk modulus K is represented as a function of pressure and temperature. The variable chamber volume is calculated based on the piston area A and the time-dependent piston position y(t) derived above. The outgoing and incoming volume flows consist of the leakage volume flow V˙ leak that flows into the gap and the volume flow caused by the piston speed (ẏ (t)A). eqn. 10 ( ) = 0 2 + � [ cos ( 2 ) + sin ( 2 )] 10 =1 = ( ) = 0 , 2 + �� , cos ( 2 ) + , sin ( 2 ) � 10 =1 1 = ̇ ( ) = 2 � �− , sin ( 2 ) + , cos ( 2 ) � 10 =1 = ( , ) ( ) � ̇ = ( , ) 0 − ( ) � ̇ ( ) − ̇ � Piston position, piston velocity, and high-pressure boundary conditions are calculated for every time step. Fluid properties Due to the extreme operating conditions mentioned above, high-pressure differences occur in the gap between the piston and the bushing, ranging from low pressure to maximum rail pressure. Accordingly, pressure drops of up to 3000 bar are possible. The same applies to the fluid temperature. While the fuel is usually supplied at temperatures between 40 and 60 °C, viscous friction in the gap can cause the fuel to heat up to approximately 200 °C. The influence of high-pressure and temperature gradients on the fluid properties is illustrated for an exemplary diesel fuel, as specified in EN 590, in Figure 3. [2, 18] As can be seen, the viscosity varies by several orders of magnitude in the temperature and pressure range shown. Across the applied characteristic map, the differences between the minimum and maximum values are still in the range of several tens of percent. The values determined for diesel show that the fluid mechanical properties necessary to solve the Reynolds equation and the energy equation are strongly influenced by pressure and temperature gradients. To obtain accurate results for friction and wear within the gap, the fluid properties have to be displayed as a function of pressure and temperature. Depending on the available literature, already parametrized fluid models are used, or suitable models were fitted based on measurement data to consider the influence of pressure and temperature. The fluid models used for the regression analysis are presented briefly in the following. Science and Research 35 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0005 Figure 2: kinematic system of the roller/ cam gear eqn. 13 To fit the measures of heat capacity and thermal conductivity, the exponential model with a polynomial term in both the numerator and denominator as given in eqn. 14 was applied. eqn. 14 As this contribution aims to investigate the influence of fuel composition on pump efficiency, the following mixing rules (eqn. 15 eqn. 17) are utilized to display the blend properties. eqn. 15 eqn. 16 eqn. 17 = = 1 ( , ) , ( , ) = exp � ∑ ∑ =0 =0 1 + ∑ ∑ =0 =0 � ( ) = � ∙ ( ) 1 = � = � ∙ , , = � ∙ , Science and Research 36 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0005 The dynamic viscosity η is modeled using an equation based on the Barus equation (eqn. 11). This equation consists of the Arrhenius equation to represent temperature dependence of viscosity and an additional exponential term b to represent pressure dependence. The exponent is modeled by a polynomial function that depends on both temperature and pressure. [20] eqn. 11 To display the influence of pressure and temperature on fluid density, the Tait equation from eqn. 12 is utilized. [21] eqn. 12 The bulk modulus necessary to solve the pressure buildup equation (eqn. 10) is derived directly from the density using eqn. 13. ( , ) = ∙ + ∙ ���� =0 =0 � ∙ � ( , ) = 0 ( 1 − 1 ) ∙ � + � 1 ( ) = � =0 Figure 3: fluid properties of diesel EN 590 depending on temperature and pressure [7, 18] For the dynamic viscosity η the logarithmic Arrhenius approach is used (eqn. 15). The density ρ mix of the mixture corresponds to the harmonic mean value of the individual mixture components. The influence of blend composition x i on heat capacity c p,mix and heat conductivity λ mix is described by calculating the arithmetic average (eqn. 17). Simulation model Reynolds equation (eqn. 5), energy equation (eqn. 6), as well as the equations for the piston kinematics (eqn. 8, eqn. 9) and pressure build-up (eqn. 10) are incorporated in the algorithm visible in Figure 4. The first step is to define the boundary conditions. These include the temperature on the high-pressure side T HP , the rotational speed ω of the drive shaft, the pressure p LP , and the gap height h. Based on the rotational speed ω piston position y and velocity ẏ are calculated by means of eqn. 8 and eqn. 9. The high pressure p HP is in turn calculated using the piston position y and velocity ẏ . The system of equations F is then set up, which includes the Reynolds equation and the energy equation. The two domains are coupled via the pressureand temperature-dependent fluid models described before. The system of equations can be solved using the Jacobian matrix J, which is derived from F. In this case, a Newton solver is used, which adjusts the Jacobian matrix J and function values F until the solution for the pressure and temperature field between two iteration steps is smaller than a specified residual res min . The described algorithm is repeated until the final simulation time is reached. Results The results for the pump efficiency obtained for an exemplary mixture of n-heptane and butyl levulinate (bl) are shown in Figure 5. The share of bl is visible on the x-axis, while the pressure difference is plotted on the y-axis. To compare the efficiency of the pump with the injection characteristics, constant Weber and Reynolds numbers are indicated by blue and red lines. When comparing the pump efficiency Φ, it can be observed that at low pressures, the highest efficiency is achieved with pure n-heptane (0 % bl). Increasing the pressure shifts the optimum mixture composition along the dashed line toward higher bl proportions. For pressures higher than 2000 bars, the highest efficiency is achieved with 100 % bl. To maintain a constant Weber number in the mixture system, the injection pressure must be increased, as the higher surface tension of bl has to be overcome. This reduces efficiency due to volumetric and hydromechanical Science and Research 37 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0005 Figure 4: simulation algorithm Symbol Description A surface area a i , b i general coefficients α d nozzle coefficient β thermal expansion coefficient c p heat capacity d nozzle diameter F system of equations h gap height H calorific value HP high pressure J Jacobian matrix K bulk modulus LP low pressure n rotational speed p pressure P mech mechanical power Q˙ heat energy flux r radius Re Reynolds number res residual t time T temperature u,v velocity components V volume V˙ volume flow We l Weber number w mass fraction x,y,z coordinates ẏ piston velocity Δp pressure difference η dynamic viscosity η hm hydromechanical efficiency η vol volumetric efficiency Science and Research 38 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0005 losses, so that up to 2400 bar pure n-heptane is most efficient (dash-dotted line). Only for very high pressures does a higher bl ratio improve efficiency Φ while keeping the Weber number constant. Summary and Outlook As part of this contribution, initial results on pump efficiency were obtained using the simulation model described in the previous section. The results show that, for binary fuel mixtures, an optimal composition can be found depending on the operating point and mixture composition. However, parameters such as elastic deformation and heat conduction in solids, which also have a major influence on leakage and friction, were neglected. The same applies to friction in the pump’s journal bearings, which was also not taken into account. For this reason, the existing model is being expanded to include the aforementioned influences, enabling a more accurate characterization of the losses and pump efficiency. The dimensionless Weber and Reynolds numbers are simplified measures that can be used to qualitatively estimate the injection behavior. In the future, therefore, more accurate methods for describing injection behavior, which have already been developed at the Fuel Science Center, will be incorporated into the model. Acknowledgement The Fuel & Chemical Science Center (FSC 2 ) is funded by the German Council of Science and Humanities (Wissenschaftsrat, WR) and by Deutsche Forschungsgemeinschaft (DFG). Figure 5: Injection pump efficiency for a mixture of n-heptane and butyl levulinate λ thermal conductivity ρ l liquid density σ surface tension ϑ temperature Φ pump efficiency ω angular velocity References [1] Robert Bosch GmbH: Handbook of diesel engines: With 86 tables. Berlin, Heidelberg: Springer, 2010 [2] Heitzig, Stefan: Analyse und Optimierung biokraftstoffgeschmierter Tribosysteme in Common-Rail-Pumpen. Dissertation, 2017 (Reihe Fluidtechnik D Band 88) [3] Rückert, Marcel; Murrenhoff, Hubertus; Heitzig, Stefan: Influence of a New Hollow Piston Design on Volumetric Losses of a Common-Rail Pump. In: ASME (Hrsg.): Proceedings of the ASME/ BATH Symposium on Fluid Power and Motion Control - 2017: Presented at ASME/ BATH 2017 Symposium on Fluid Power and Motion Control, October 16-19, 2017, Sarasota, Florida, USA. New York, N.Y.: The American Society of Mechanical Engineers, 2017 [4] Rückert, Marcel; Heitzig, Stefan; Murrenhoff, Hubertus: Reducing Friction and Wear in Common-Rail Pumps by Grooved and Contoured Pistons. In: ASME (Hrsg.): Proceedings of the BATH/ ASME Symposium on Fluid Power and Motion Control - 2016: Presented at BATH/ ASME 2016 Symposium on Fluid Power and Motion Control, September 7-9, 2016, Bath, United Kingdom. New York, N.Y.: The American Society of Mechanical Engineers, 2016 [5] Weinebeck, Alexander; Heitzig, Stefan; Murrenhoff, Hubertus: Tribological Investigations of new biofuel blends. In: Bartz, Wilfried J. (Hrsg.): Tribology - industrial and automotive lubrication: 19 th International Colloquium Tribology, 21 - 23 January 2014, [Stuttgart/ Ostfildern. Esslingen, Ostfildern: TAE, 2014 [6] Hofmeister, Marius; Fischer, Felix Justin; Boden, Lukas; Schmitz, Katharina: Challenges in the Use of Bio-Hybrid Fuels As Drop-in Fuels. In: ASME (Hrsg.): Proceedings of ASME 2024 18th International Conference on Energy Sustainability (ES2024): July 15-17, 2024, Anaheim, California. Little Falls, NJ, USA, Red Hook, NY: The American Society of Mechanical Engineers; Curran Associates Inc, 2024 [7] Europäisches Komitee für Normung: EN 590: 2013+A1: 2017. A1: 2017. Brüssel: CEN, 2017 (Dieselkraftstoff - Anforderungen und Prüfverfahren) [8] International Organization for Standardization: Road vehicles - Calibration fluids for diesel injection equipment. Geneva: ISO, 2010 (ISO 4113) [9] International Organization for Standardization: Diesel fuel - Assessment of lubricity using the high-frequency reciprocating rig (HFRR). Geneva: ISO, 2018 (ISO 12156-1) [10] Ackermann, Philipp; Braun, Karsten E.; Burkardt, Patrick; Heger, Sebastian; König, Andrea; Morsch, Philipp; Lehrheuer, Bastian; SURGER, Maximilian; Völker, Simon; Blank, Lars Mathias; Du, Miaomiao; Heufer, Karl Alexander; Roß-Nickoll, Martina; Viell, Jörn; Aßen, Niklas von der; Mitsos, Alexander; Pischinger, Stefan; Dahmen, Manuel: Designed to Be Green, Economic, and Efficient: A Ketone-Ester-Alcohol-Alkane Blend for Future Spark-Ignition Engines. In: ChemSusChem 14 (2021), Nr. 23, S. 5254-5264 [11] Grütering, Carolin; Honecker, Christian; Hofmeister, Marius; Neumann, Marcel; Raßpe-Lange, Lukas; DU, Miaomiao; Lehrheuer, Bastian; Campenhausen, Maximilian von; Schuster, Franziska; Surger, Maximilian; Ebert, Birgitta E.; Jupke, Andreas; Tiso, Till; Leonhard, Kai; Schmitz, Katharina; Pischinger, Stefan; Blank, Lars M.: Methyl ketones: a comprehensive study of a novel biofuel. In: Sustainable Energy & Fuels 8 (2024), Nr. 9, S. 2059- 2072 [12] Voelker, Simon; Deutz, Sarah; Burre, Jannik; Bongartz, Dominik; Omari, Ahmad; Lehrheuer, Bastian; Mitsos, Alexander; Pischinger, Stefan; Bardow, André; Assen, Niklas von der: Blend for all or pure for few? Well-to-wheel life cycle assessment of blending electricity-based OME 3-5 with fossil diesel. In: Sustainable Energy & Fuels 6 (2022), Nr. 8, S. 1959-1973 [13] Voelker, Simon; Groll, Niklas; Bachmann, Marvin; Mueller, Leonard; Neumann, Marcel; Kossioris, Theodoros; Muthyala, Paul; Lehrheuer, Bastian; Hofmeister, Marius; Vorholt, Andreas; Schmitz, Katharina; Pischinger, Stefan; Leitner, Walter; Bardow, André: Towards carbon-neutral and clean propulsion in heavy-duty transportation with hydroformylated Fischer-Tropsch fuels. In: Nature Energy 9 (2024), Nr. 10, S. 1220-1229 [14] Charalampous, Georgios; Hardalupas, Yannis: How do liquid fuel physical properties affect liquid jet development in atomisers? In: Physics of Fluids 28 (2016), Nr. 10 [15] Chu, Hongchao; Welch, Cooper; Elmestikawy, Hani; Cao, Shangyi; Davidovic, Marco; Böhm, Benjamin; Dreizler, Andreas; Pitsch, Heinz: A Combined Numerical and Experimental Investigation of Cycle-to-Cycle Variations in an Optically Accessible Spark-Ignition Engine. In: Flow, Turbulence and Combustion 110 (2023), Nr. 1, S. 3-29 [16] Deshmukh, Abhishek Y.; Davidovic, Marco; Grenga, Temistocle; Lakshmanan, Raghavan; Cai, Liming; Pitsch, Heinz: A reduced-order model for turbulent reactive sprays in compression ignition engines. In: Combustion and Flame 236 (2022), S. 111751 [17] Fischer, Felix Lukas: Thermoelastohydrodynamische Betrachtung des Kolben-Buchse-Kontakts in Hochdruckpumpen. Dissertation, 2020 (Reihe Fluidtechnik D Band 105) [18] Drumm, Sebastian Michael: Entwicklung von Messmethoden hydraulischer Kraftstoffeigenschaften unter Hochdruck. Zugl.: Aachen, Techn. Hochsch., Diss., 2012. Aachen: Shaker, 2012 (Reihe Fluidtechnik D 69) [19] Bartel, Dirk: Simulation von Tribosystemen: Grundlagen und Anwendungen. Zugl.: Magdeburg, Univ., Fak. für Maschinenbau, Habil.-Schr., 2009. 1. Aufl. Wiesbaden: Vieweg + Teubner, 2010 (Vieweg + Teubner Research) [20] Hannon, W. M.; Hager, C. H.; Sadinski, R. J.: A Comparison of the High-Pressure Viscometry and Optical Inference Methods Used to Determine the Pressure-Viscosity Coefficient. In: Journal of Tribology 145 (2023), Nr. 5 [21] Dymond, J. H.; Malhotra, R.: The Tait equation: 100 years on. In: International Journal of Thermophysics 9 (1988), Nr. 6, S. 941-951 Science and Research 39 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 DOI 10.24053/ TuS-73-0005 Science and Research 40 Tribologie + Schmierungstechnik · volume 73 · issue 1/ 2026 MEDIENTIPP Erik Kuhn On the Tribology of Lubricating Greases An energetic approach to post-modern tribology Tribologie - Schmierung, Reibung, Verschleiß 1. Auflage 2025, 204 Seiten ISBN print 978-3-381-14171-5 ISBN eBook 978-3-381-14172-2 DOI 10.24053/ 9783381141722 Ladenpreis print €[D] 118,00 Ladenpreis eBook €[D] 94,99 This monograph takes a new look at tribology with its basic concepts of friction and wear using the example of lubricating greases. The consideration of the phenomenon of occurring instabilities and the introduction of the entropy concept into lubricating grease tribology provide a new perspective on known phenomena. The second part of this book presents a wide range of experimental possibilities for investigating lubricating greases. Contents Introduction to Instability and Postmodern Tribology - On the Phenomenon of Self - Organization - Postmodern Grease Tribology - Lubricating Grease - Rheological behavior of Lubricating greases - A Selected Traditional Wear Model - The Extension of the Wear Concept expert verlag - Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. 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You can obtain the full open access service for a one-off article processing charge of € 1,850.00 (plus VAT). Editor in chief Dr. Manfred Jungk eMail: jungk@verlag.expert Publisher expert verlag Ein Unternehmen der Narr Francke Attempto Verlag GmbH + Co. KG Dischingerweg 5 D-72070 Tübingen Tel.: +49 (0)7071 97 556 0 eMail: info@verlag.expert www.expertverlag.de Editorial office Angelika Böer eMail: boeer@narr.de 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 Jakob Gleichner, Markus Golek Development of a squeeze film damper test rig - comparison of measurement and simulation results Felix Leuf, Georg Jacobs, Jan Euler, Patrick Laubichler, Michael Mirza, Thomas Decker, Julian Röder Evaluation of the start-up behavior of a novel spherical plain bearing as a main bearing for wind turbines based on scaled tests Yongzhen Lin, Matthias Kröger, Jenny Köckritz, Deeksha Lohani, Ringo Nepp Tribological investigations of radial shaft seal with focus on 3D printed counter surface Marius Hofmeister, Katharina Schmitz Increasing efficiency of diesel injection pumps by targeted blending of bio-hybrid fuels