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