Browsing by Subject "bearings"
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Item An Experimental Investigation of the Static and Dynamic Performance of Horizontal-Application and Vertical-Application Three-Lobe Bearings(2014-08-07) Khatri, RasishStatic and dynamic performance test results are provided for a horizontal-application three-lobe bearing evaluated over the following range of static-load orientations (all taken from the leading edge of the loaded pad): 0?, 20?, 30?, 40?, 60?, 80?, 90?, and 100?. This bearing has the following specifications: 100? pad arc angle, 0.52 preload, 70% offset, 101.74 mm (4.0057 in) minimum bore diameter, 0.116 mm (7.55 mils) radial pad clearance, and 76.3 mm (3 in) axial length. The static and dynamic test results are evaluated to determine the sensitivity of the bearing to changes in the static load direction. The two questions the study aims to answer are: (1) ?Is an offset three-lobe bearing a good choice when the radial static load vector represents an unknown variable?? and (2) ?Can an offset three-lobe bearing be oriented advantageously with a known load direction?? Both the static and dynamic test results are compared to predictions obtained from a fixed-arc bearing Reynolds equation solver. Predictions using both the measured hot clearance and measured cold clearance as inputs are compared to the measured data. Dynamic tests show that the horizontal-application three-lobe bearing is sensitive to load orientation at low speeds and high loads. Whirl-frequency ratios (WFR) at 6750 rpm with loads of 1149 kPa, 1723 kPa, and 2298 kPa are equal to zero for loads oriented towards the leading edge of the pad, and between 0.35 and 0.5 for loads oriented towards the trailing edge of the pad. This same general trend can be seen for WFR values at 9000 rpm and 10800 rpm. The horizontal-application three-lobe bearing is not sensitive to load orientation at high speeds and light loads. At 13200 rpm, measured WFRs are between 0.3 and 0.7 at all loads and for all load orientations. Measured WFR at the no-load condition are between 0.45 and 0.7 for all cases. Stiffness orthotropy was found to vary significantly with load orientation. At 6750 rpm and 2298 kPa, the bearing is most orthotropic when the static load orientation is 30? and 40?, with K_(yy) being larger than K_(xx) by approximately 800 MN/m. At 13200 rpm and 2298 kPa, the bearing is most orthotropic for the 40? and 100? load orientations, with K_(yy) being larger than K_(xx) by approximately 800 MN/m. Overall, it is concluded that the three-lobe bearing is not a good choice when the load direction is unknown, as the bearing can have different rotordynamic coefficients with different load orientations. Also, at high speeds, the three-lobe bearing cannot be oriented advantageously with a known load direction to enhance stability, as the WFR of the three-lobe bearing tested is largely independent of load direction. However, the three-lobe bearing can be oriented advantageously to split/change a critical speed, since the direct stiffness and stiffness orthotropy change with load direction. Additionally, dynamic performance test results are provided for a vertical-application (nominally unloaded) three-lobe bearing. The vertical-application bearing has the following specifications: 100? pad arc angle, 0.64 preload, 100% offset, 101.74 mm (4.0057 in) nominal diameter, 0.116 mm (5.27 mils) radial pad clearance, 76.3 mm (3in) axial length, and 100? static load orientation from the leading edge of the loaded pad. The performance of this bearing is evaluated to determine: (1) whether a fully (100%) offset three-lobe bearing configuration is more stable in terms of the WFR than a standard plain journal bearing and (2) whether a fully offset three-lobe bearing provides a larger direct stiffness (centering force) than a standard fixed-arc bearing. Dynamic tests show that the vertical-application three-lobe bearing does not improve stability over conventional plain journal bearings. The measured WFRs for the vertical-application bearing are approximately 0.4-0.5 for nearly all test cases. Predicted WFRs are 0.46 at all test points. The vertical-application bearing dimensionless direct stiffness coefficients are compared to those for the horizontal-application bearing. The equivalent stiffness for the vertical-application bearing is larger than that of the horizontal-application bearing by a factor of 1.33 at 6750 rpm and a factor of 1.25 at 9000 rpm. Thus, the vertical-application bearing does impart a larger centering force to the journal relative to the horizontal-application bearing when the journal is not carrying a radial static load.Item Experimental frequency-dependent rotordynamic coefficients for a load-on-pad, high-speed, flexible-pivot tilting-pad bearing(Texas A&M University, 2004-09-30) Rodriguez Colmenares, Luis EmigdioThis thesis provides experimental frequency dependent stiffness and damping coefficient results for a high-speed, lightly loaded, flexible-pivot tilting-pad bearing, with a load-on-pad configuration. Test conditions include four shaft speeds (6000, 9000, 13000 and 16000 rpm), and bearing unit loads from 172 kPa to 690 kPa. The results show that the bearing stiffness is a quadratic function of the frequency of vibration; hence their frequency dependency can be modeled by added-mass terms. The additional degrees of freedom introduced by the pads and the influence of the inertial forces generated in the fluid film account for this frequency dependency. The conventional frequency-dependent stiffness and damping model for tilting-pad bearings is extended with an added-mass matrix to account for the frequency dependency. This approach allows the description of the bearing dynamic characteristics with frequency-independent stiffness, damping and added-mass matrices. Experimental results are compared with predictions from the Reynolds equation and from a bulk-flow Navier-Stokes model. Both models produce good predictions of the stiffness and damping coefficients. However, results show that the bulk-flow model is more adequate for predicting the direct added-mass terms because it accounts for the fluid inertial forces. A bulk-flow solution of the Navier-Stokes equations that includes the effects of fluid inertia should be used to calculate the rotordynamic coefficients of a flexible-pivot tilting-bearing. Static performance measurement results are also detailed. Results include pad metal temperatures, eccentricity-ratios and attitude-angle as a function of bearing load, and estimated power losses.Item High Temperature, Permanent Magnet Biased Magnetic Bearings(2010-07-14) Gandhi, Varun R.The Electron Energy Corporation (EEC) along with the National Aeronautics and Space Administration (NASA) is researching magnetic bearings. The purpose of this research was to design and develop a high-temperature (1000?F) magnetic bearing system using High Temperature Permanent Magnets (HTPM), developed by the EEC. The entire system consisted of two radial bearings, one thrust bearing, one motor and 2 sets of catcher bearings. This high temperature magnetic bearing system will be used in high performance, high speed and high temperature applications like space vehicles, jet engines and deep sea equipment. The bearing system had a target design to carry a load equal to 500 lb-f (2225N). Another objective was to design and build a test rig fixture to measure the load capacity of the designed high temperature radial magnetic bearing (HTRMB) called Radial Bearing Force Test Rig (RBFTR). A novel feature of this high temperature magnetic bearing is its homopolar construction which incorporates state of the art high temperature, 1000 ?F, permanent magnets. A second feature is its fault tolerance capability which provides the desired control forces even if half the coils have failed. The permanent magnet bias of the radial magnetic bearing reduces the amount of current required for magnetic bearing operation. This reduces the power loss due to the coil current resistance and also increases the system efficiency because magnetic field of the HTPM is used to take up the major portion of the static load on the bearing. The bias flux of the homopolar radial bearing is produced by the EEC HTPM to reduce the related ohmic losses of an electromagnetic circuit significantly. An experimental procedure was developed using the Radial Bearing Force Test Rig (RBTFR) to measure actual load capacity of the designed bearing at the test rig. All the results obtained from the experiment were compiled and analyzed to determine the relation between bearing force, applied current and temperature.Item Identification of force coefficients in flexible rotor-bearing systems - enhancements and further validations(Texas A&M University, 2005-11-01) Balantrapu, Achuta Kishore Rama KrishnaRotor-bearing system characteristics, such as natural frequencies, mode shapes, stiffness and damping coefficients, are essential to diagnose and correct vibration problems during system operation. Of the above characteristics, accurate identification of bearing force parameters, i.e. stiffness and damping coefficients, is one of the most difficult to achieve. Field identification by imbalance response measurements is a simple and often reliable way to determine synchronous speed force coefficients. An enhanced method to estimate bearing support force coefficients in flexible rotor-bearing systems is detailed. The estimation is carried out from measurements obtained near bearing locations from two linearly independent imbalance tests. An earlier approach assumed rotordynamic measurements at the bearing locations, which is very difficult to realize in practice. The enhanced method relaxes this constraint and develops the procedure to estimate bearing coefficients from measurements near the bearing locations. An application of the method is presented for a test rotor mounted on two-lobe hydrodynamic bearings. Imbalance response measurements for various imbalance magnitudes are obtained near bearing locations and also at rotor mid-span. At shaft speeds around the bending critical speed, the displacements at the rotor mid-span are an order of magnitude larger than the shaft displacements at the bearing locations. The enhanced identification procedure renders satisfactory force coefficients in the rotational speed range between 1,000 rpm and 4,000 rpm. The amount of imbalance mass needed to conduct the tests and to obtain reliable shaft displacement measurements influences slightly the magnitude of the identified force coefficients. The effect of increasing the number of rotor sub-elements in the finite-element modeling of the shaft is noted. Sensitivity of the method and derived parameters to noise in the measurements is also quantified.