Squeeze film dampers (SFDs) in aircraft engines effectively aid to reduce rotor motion amplitudes, in particular when traversing a critical speed, and help to alleviate rotor whirl instabilities. The current work is a long-term endeavor focused on quantifying the dynamic force performance of practical SFDs, exploring novel design damper configurations, and producing physically sound predictive SFD models validated by experimental data. Piston rings (PRs) and O-rings (ORs), commonly used as end seals in SFDs for commercial and military gas turbine engines, respectively, amplify viscous damping in a short physical length and while operating with a modicum of lubricant flow. This paper presents experimental force coefficients (damping and inertia) for two identical geometry SFDs with end seals, one configuration hosts PRs, and the other one ORs. The test rig comprises a stationary journal and bearing cartridge (BC) hosting the SFD and supported on four elastic rods to emulate a squirrel cage. The damper film land length, diameter, and clearance are L = 25.4 mm, D = 5L, and c = 0.373 mm (D/c = 340), respectively. A supply feeds ISO VG 2 oil to the film land at its middle plane through either one hole or three holes, 2.5 mm in diameter, 120 deg apart. In the PRSFD, the lubricant exits through the slit opening at the ring butted ends. The ORs suppress oil leakage; hence, lubricant evacuates through a 1 mm hole at ¼ L near one journal end. The ORs when installed add significant stiffness and damping to the test structure. The ORSFD produces 20% more damping than the PRSFD, whereas both sealed ends SFDs show similar size added mass. For oil supplied at 0.69 bar(g) through a single orifice produces larger damping, 60–80% more than when the damper operates with three oil feedholes. A computational model reproducing the test conditions delivers force coefficients in agreement with the test data. Archival literature calls for measurement of a single pressure signal to estimate SFD reaction forces. For circular centered orbits (CCOs), the dynamic pressure field, in the absence of any geometrical asymmetry or feed/discharge oil condition, “rotates” around the bearing with a speed equal to the whirl frequency. The paper presents force coefficients estimated from (a) measurements of the applied forces and ensuing displacements, and (b) the dynamic pressure recorded at a fixed angular location and “integrated” over the journal surface. The first method delivers a damping coefficient that is large even with lubricant supplied at a low oil supply pressure whereas the inertia coefficient increases steadily with feed pressure. Predictions show good agreement with the test results from measured forces and displacements, in particular the added mass. On the other hand, identified damping and inertia coefficients from dynamic pressures show a marked difference from one pressure sensor to another, and vastly disagreeing with test results from the first method or predictions. The rationale for the discrepancy relies on local distortions in the dynamic pressure fields that show zones of oil vapor cavitation at a near zero absolute pressure and/or with air ingestion producing high frequency spikes from bubble collapsing; both phenomena depend on the magnitude of the oil supply pressure. An increase in lubricant supply pressure suppresses both oil vapor cavitation and air ingestion, which produces an increase of both damping and inertia force coefficients. No prior art compares the performance of a PRSFD vis-à-vis that of an ORSFD. Supplying lubricant with a large enough pressure (flow rate) is crucial to avoid the pervasiveness of air ingestion. Last, the discussion on force coefficients obtained from two distinct methods questions the use of an oversimplifying assumption; the dynamic pressure field is not invariant in a rotating coordinate frame.

High-performance turbomachinery demands high shaft speeds, increased rotor flexibility, tighter clearances in the flow passages, advanced materials, and increased tolerance to imbalances. Operation at high speeds induces severe dynamic loading with large amplitude shaft motions at the bearing supports. Typical rotordynamic models rely on linearized force coefficients (stiffness K, damping C, and inertia M) to model the reaction forces from fluid film bearings and seal elements. These true linear force coefficients are derived from infinitesimally small amplitude motions about an equilibrium position. Often, however, a rotor–bearing system does not reach an equilibrium position and displaces with motions amounting to a sizable portion of the film clearance; the most notable example being a squeeze film damper (SFD). Clearly, linearized force coefficients cannot be used in situations exceeding its basic formulation. Conversely, the current speed of computing permits to evaluate fluid film element reaction forces in real time for ready numerical integration of the transient response of complex rotor–bearing systems. This approach albeit fast does not help to gauge the importance of individual effects on system response. Presently, an orbit analysis method estimates force coefficients from numerical simulations of specified journal motions and predicted fluid film reaction forces. For identical operating conditions in static eccentricity and whirl amplitude and frequency as those in measurements, the computational physics model calculates instantaneous damper reaction forces during one full period of motion and performs a Fourier analysis to characterize the fundamental components of the dynamic forces. The procedure is repeated over a range of frequencies to accumulate sets of forces and displacements building mechanical transfer functions from which force coefficients are identified. These coefficients thus represent best fits to the motion over a frequency range and dissipate the same mechanical energy as the nonlinear mechanical element. More accurate than the true linearized coefficients, force coefficients from the orbit analysis correlate best with SFD test data, in particular for large amplitude motions, statically off-centered. The comparisons also reveal the fallacy in representing nonlinear systems as simple K–C–M models impervious to the kinematics of motion.

Aircraft engines customarily implement squeeze film dampers (SFDs) to dissipate mechanical energy caused by rotor vibration and to isolate the rotor from its structural frame. The paper presents experimental results for the dynamic forced performance of an open ends SFD operating with large amplitude whirl motions, centered and off-centered. The test rig comprises of an elastically supported bearing with a damper section, 127 mm in diameter, having two parallel film lands separated by a central groove. Each film land is 25.4 mm long with radial clearance c = 0.251 mm. The central groove, 12.7 mm long, has a depth of 9.5 mm (38c). An ISO VG 2 lubricant flows into the groove via three 2.5 mm orifices, 120 deg apart, and then passes through the film lands to exit at ambient condition. Two orthogonally placed shakers apply dynamic loads on the bearing to induce circular orbit motions with whirl frequency ranging from 10 Hz to 100 Hz. A static loader, 45 deg away from each shaker, pulls the bearing to a static eccentricity (e_s). Measurements of dynamic loads and the ensuing bearing displacements and accelerations, as well as the film and groove dynamic pressures, were obtained for eight orbit amplitudes (r = 0.08c to ∼0.71c) and under four static eccentricities (e_s = 0.0c to ∼0.76c). The experimental damping coefficients increase quickly as the bearing offset increases (e_s/c → 0.76) while remaining impervious to the amplitude of whirl orbit (r/c → 0.51). The inertia coefficients decrease rapidly as the orbit amplitude grows large, r > 0.51c, but increase with the static eccentricity. A comparison with test results obtained with an identical damper but having a smaller clearance (c_s = 0.141 mm) (San Andrés, L., 2012, “Damping and Inertia Coefficients for Two Open Ends Squeeze Film Dampers With a Central Groove: Measurements and Predictions,” ASME J. Eng. Gas Turbines Power, 134(10), p. 102506), show the prior damping and inertia coefficients are larger, ∼5.0 and ∼2.2 times larger than the current ones. These magnitudes agree modestly with theoretical ratios for damping and inertia coefficients scaling as (c/c_s)³= 5.7 and (c/c_s) = 1.8, respectively. In spite of the large difference in depths between a groove and a film land, the magnitudes of dynamic pressures recorded at the groove are similar to those in the lands. That is, the groove profoundly affects the dynamic forced response of the test damper. A computational physics model replicates the experimental whirl motions and predicts force coefficients spanning the same range of whirl frequencies, orbit radii, and static eccentricities. The model predictions reproduce with great fidelity the experimental force coefficients. The good agreement relies on the specification of an effective groove depth derived from one experiment.

Squeeze film dampers (SFDs) are effective to ameliorate shaft vibration amplitudes and to suppress instabilities in rotor–bearing systems. Compact aero jet engines implement ultra-short length SFDs (L/D ≤ 0.2) to satisfy stringent weight and space demands with low parts count. This paper describes a test campaign to identify the dynamic forced response of an open ends SFD (L = 25.4 mm and D = 125.7 mm), single film land, and oil fed through three holes (120 deg apart), operating with similar conditions as in an aircraft engine. Two journals make for two SFD films with clearances c_A = 0.129 mm and c_B = 0.254 mm (small and large). The total oil-wetted length equals L_tot = 36.8 mm that includes deep end grooves, width and depth = 2.5 × 3.8 mm, for installation of end seals. In the current experiments, the end seals are not in place. A hydraulic static loader pulls the bearing cartridge (BC) to a preset static eccentricity (e_S), and two electromagnetic shakers excite the BC with single frequency loads to create circular orbits, centered and off-centered, over a prescribed frequency range ω = 10–100 Hz. The whirl amplitudes range from r = 0.05c_A–0.6c_A and r = 0.15c_B–0.75c_B while the static eccentricity increases to e_S = 0.5c_A and e_S = 0.75c_B, respectively. Comparisons of force coefficients between the two identical dampers with differing clearances show that the small clearance damper (c_A) provides ∼4 times more damping and ∼1.8 times the inertia coefficients than the damper with large clearance (c_B). The test results demonstrate damping scales with ∼1/c³ and inertia with ∼1/c, as theory also showed. Analysis of the measured film land pressures evidence that the deep end grooves contribute to the generation of dynamic pressures enhancing the dynamic forced response of the test SFDs. A thin film flow model with an effective groove depth delivers predictions that closely match the test damping and inertia coefficients. Other predictions, based on the short length bearing model, use an effective length L_eff ∼ 1.17L to deliver damping coefficients 15% larger than the experimental results; however, inertia coefficients are ½ of the identified magnitudes. The experiments and analysis complement earlier experimental work conducted with centrally grooved SFDs.

This paper presents the change of non-dimensional characteristics and thermal behavior of different sized tilting pad journal bearings (TPJBs) with the same Sommerfeld number. A three-dimensional (3D) TPJB numerical model is provided considering the thermo-elastic hydro-dynamic (TEHD) lubrication model with pad thermal-elastic deformation. The pivot stiffness is assumed to be the combination of linear and cubic stiffness based on the Hertzian contact theory. The TPJBs in a configuration of load between pad (LBP) with the same Sommerfeld number having seven different sizes are simulated, and their non-dimensional dynamic and static characteristics and thermal behavior are compared. Pad thermal and elastic deformation are both taken into account. If the changes in lubricant viscosity, thermal deformation, and elastic deformation of journal/pads due to viscous shearing are ignored, the bearings with identical Sommerfeld numbers should show the same performance characteristics. However, the heat generation at the bearing clearance during operation (a) induces a decrease in viscosity and heat transfer to journal/pads and (b) results in a thermal deformation. Furthermore, the elastic deformation of the tilting pads and pivots also affects the bearing dynamic performance. For the same Sommerfeld number, the numerical analyses provide how the viscous shearing and elastic deformation lead to a change in bearing performance. For the small bearings with the same Sommerfeld number, the non-dimensional characteristics did not change significantly, where the heat generation was small being compared to the large sized bearing. The largest change in non-dimensional characteristics occurred when the maximum temperature of the oil film increased by 30 °C or more compared to the lubricant supply temperature. The root cause of the change in the non-dimensional characteristics is the viscous shearing in the oil film, and the thermal deformation of the structures surrounding the oil film acts in combination. These results provide insight into the Sommerfeld number, which can be used for the early stage of bearing design.

Squeeze film dampers (SFDs) aid to both reduce rotor dynamic displacements and to increase system stability. Dampers sealed with piston rings (PR), common in aircraft engines, are proven to boost damping generation, reduce lubricant flow demand, and prevent air ingestion. This paper presents the estimation of force coefficients in a short length SFD, PR sealed, and supplied with a light lubricant at two feed pressures, P_in-1~0.69 barg and P_in-2~2.76 barg, i.e., low and high. Two pairs of PRs are installed in the test SFD, one set has flow conductance C_S1 = 0.56 LPM/bar, whereas the other pair has C_S2 = 0.89 LPM/bar. The second set leaks more as it has a larger slit gap. Dynamic load tests show that both dampers, having seal flow conductances differing by 60%, produce damping and added mass coefficients of similar magnitude, differing by at most 20%. Other experiments quantify the effect of lubricant supply pressure, P_in-1 and P_in-2, on the dynamic film pressure and force coefficients of the PR-SFD. The damper configuration with C_S1 and operating with the high P_in-2 shows ~20% more damping and added mass coefficients compared to test results for the damper supplied with P_in-1. Film pressure measurements show that the air ingestion and oil vapor cavitation coexist for operation at the low P_in-1. Computational predictions accounting for the feed holes in the physical model agree with the experimental coefficients. On the other hand, predictions from classical formulas for an idealized damper geometry, fully sealed at its ends, largely over predict the measured force coefficients.