Bearing Nominal Dimensions) 30 Predicted Structural Stiffness of a Single Bump for Different Dry Friction Coefficients. Foil Bearing Nominal Dimensions 31 Predicted Structural Stiffness of a Single Bump for a ± 30% Range of the Nominal Bump Height. (µ f = 0.1; Foil Bearing Nominal Dimensions) 32 Predicted Structural Stiffness of a Single Bump for a ± 10% Range of the Nominal Half Bump Length. (µ f = 0.1; Foil Bearing Nominal Dimensions) 33 Coordinate Systems in the CFB for the Structural Stiffness Analysis 34 Predicted and Experimental Load versus Bearing Deflection Curves for Position 1-5 (0º-180º). (Shaft Diameter D 1 = 1.500 in; µ f = 0.1; lo = 0.08 in) 35 Predicted and Experimental Load versus Bearing Deflection Curves for Position 1-5 (0º-180º). (Shaft Diameter D 1 = 1.500 in; µ f = 0.1; lo = 0.085 in) 36 Predicted and Experimental Load versus Bearing Deflection Curves for Position 1-5 (0º-180º). (Shaft Diameter D 1 = 1.500 in; µ f = 0.1; lo = 0.07 in) 37 Predicted Load versus Bearing Deflection Curves for Different Values of Dry Friction Coefficients. (Positions 1-5; Shaft Diameter D1; lo = 0.08 in) 38 Predicted and Experimental Structural Stiffness versus the Shaft Deflection (Shaft Diameter D 1 =1.500in; µ f = 0.1; lo = 0.085 in; r = 0 in) 39 Predicted and Experimental Structural Stiffness versus the Shaft Deflection (Shaft Diameter D 2 =1.501 in; µ f = 0.1; lo = 0.085 in; r = 0.5 mil) 40 Predicted and Experimental Structural Stiffness versus the Shaft Deflection (Shaft Diameter D 3 =1.499 in; µ f = 0.1; lo = 0.085 in; r =-0.5 mil) 41 Predicted and Experimental Structural Stiffness with respect to the Static Load.
High performance oil-free turbomachinery implements gas foil bearings (FBs) to improve mechanical efficiency in compact units. FB design, however, is still largely empirical due to their mechanical complexity. The paper provides test results for the structural parameters in a bump-type foil bearing. The stiffness and damping (Coulomb or viscous type) coefficients characterize the bearing compliant structure. The test bearing, 38.1 mm in diameter and length, consists of a thin top foil supported on bump-foil strips. A prior investigation identified the stiffness due to static loads. Presently, the test FB is mounted on a non-rotating stiff shaft and a shaker exerts single frequency loads on the bearing. The dynamic tests are conducted at shaft surface temperatures from 25 °C to 75°C. Time and frequency domain methods are implemented to determine the FB parameters from the recorded periodic load and bearing motions. Both methods deliver identical parameters. The dry friction coefficient ranges from 0.05 to 0.20, increasing as the amplitude of load increases. The recorded motions evidence a resonance at the system natural frequency, i.e. null damping. The test derived equivalent viscous damping is inversely proportional to the motion amplitude and excitation frequency. The characteristic stick-slip of dry friction is dominant at small amplitude dynamic loads leading to a hardening effect (stiffening) of the FB structure. The operating temperature produces shaft growth generating a bearing preload. However, the temperature does not affect significantly the identified FB parameters, albeit the experimental range was too small considering the bearings intended use in industry.
Gas foil bearings (FB) satisfy many of the requirements noted for novel oil-free turbomachinery. However, FB design remains largely empirical, in spite of successful commercial applications. The mechanical structural characteristics of foil bearings, namely stiffness and damping, have been largely ignored in the archival literature. Four commercial bump-type foil bearings were acquired to measure their load capacity under conditions of no shaft rotation. The test bearings contain a single Teflon coated foil supported on 25 bumps. The nominal radial clearance is 0.036 mm for a 38 mm journal. A simple test set up was assembled to measure the FB deflections resulting from static loads. The tests were conducted with three shafts of increasing diameter to induce a degree of preload into the FB structure. Static measurements show nonlinear FB deflections, varying with the orientation of the load relative to the foil spot weld. Loading and unloading tests evidence hysteresis. The FB structural stiffness increases as the bumps-foil radial deflection increases (hardening effect). The assembly preload results in notable stiffness changes, in particular for small radial loads. A simple analytical model assembles individual bump stiffnesses and renders predictions for the FB structural stiffness as a function of the bump geometry and material, dry-friction coefficient, load orientation, clearance and preload. The model predicts well the test data, including the hardening effect. The uncertainty in the actual clearance (gap) upon assembly of a shaft into a FB affects most the predictions.
Gas foil bearings (GFBs) satisfy the requirements for oil-free turbomachinery, i.e. simple construction and ensuring low drag friction and reliable high speed operation. However, GFBs have a limited load capacity and minimal damping, as well as frequency and amplitude dependent stiffness and damping characteristics. This paper provides experimental results of the rotordynamic performance of a small rotor supported on two bump-type GFBs of length and diameter equal to 38.10 mm. Coast down rotor responses from 25 krpm to rest are recorded for various imbalance conditions and increasing air feed pressures. The peak amplitudes of rotor synchronous motion at the system critical speed are not proportional to the imbalance introduced. Furthermore, for the largest imbalance, the test system shows subsynchronous motions from 20.5 krpm to 15 krpm with a whirl frequency at ∼ 50% of shaft speed. Rotor imbalance exacerbates the severity of subsynchronous motions, thus denoting a forced nonlinearity in the GFBs. The rotor dynamic analysis with calculated GFB force coefficients predicts a critical speed at 8.5 krpm, as in the experiments; and importantly enough, unstable operation in the same speed range as the test results for the largest imbalance. Predicted imbalance responses do not agree with the rotor measurements while crossing the critical speed, except for the lowest imbalance case. Gas pressurization through the bearings’ side ameliorates rotor subsynchronous motions and reduces the peak amplitudes at the critical speed. Post-test inspection reveal wear spots on the top foils and rotor surface.
High performance oil-free turbomachinery implements gas foil bearings (FBs) to improve mechanical efficiency in compact units. FB design, however, is still largely empirical due to its mechanical complexity. The paper provides test results for the structural parameters in a bump-type foil bearing. The stiffness and damping (Coulomb or viscous type) coefficients characterize the bearing compliant structure. The test bearing, 38.1mm in diameter and length, consists of a thin top foil supported on bump-foil strips. A prior investigation identified the stiffness due to static loads. Presently, the test FB is mounted on a non-rotating stiff shaft and a shaker exerts single frequency loads on the bearing. The dynamic tests are conducted at shaft surface temperatures from 25to75°C. Time and frequency domain methods are implemented to determine the FB parameters from the recorded periodic load and bearing motions. Both methods deliver identical parameters. The dry friction coefficient ranges from 0.05 to 0.20, increasing as the amplitude of load increases. The recorded motions evidence a resonance at the system natural frequency, i.e., null damping. The test derived equivalent viscous damping is inversely proportional to the motion amplitude and excitation frequency. The characteristic stick-slip of dry friction is dominant at small amplitude dynamic loads leading to a hardening effect (stiffening) of the FB structure. The operating temperature produces shaft growth generating a bearing preload. However, the temperature does not significantly affect the identified FB parameters, albeit the experimental range was too small considering the bearings intended use in industry.
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