A facility and apparatus are described which determine stiffness, 'damping, and added-mass rotordynamic coefficients plus steady-state operating characteristics of high speed hydrostatic journal bearings. The apparatus has a current top speed of 29800 rpm with a bearing diameter of 7.62 cm (3 in.). Purified warm water, 55°C (130°F), is used as a test fluid to achieve elevated Reynolds numbers during operation. The test-fluid pump yields a bearing maximum inlet pressure of 6.9 Mpa (1000 psi). Static load on the bearing is independently controlled and measured. Orthogonally mounted external shakers are used to excite the test stator in the direction of, and perpendicular to, the static load. The apparatus can independently calculate all rotordynamie coefficients at a given operating condition. IntroductionExperience with the SSME (Space Shuttle Main Engine) has demonstrated definite limits for the successful operation of ball bearings in liquid oxygen (LOX). Currendy, bearings in the HI_TP (High Pressure Oxygen Turbopump) of the SSME experience accelerated wear at full power level. The balls in these bearings simply get smaller, rapidly. The problem has proven largely intractable. As a result, hydrostatic bearings have been proposed for many new turbopump applications because of their long lifetime, low friction factors, low wear, and their ability to use low-viscosity lubricants. For cryogenic applications, the bearings will be pressurized from pump-discharge flow. At zero speeds the bearings will be flooded but unpressurized. Rubbing will be experienced at start up and shut down as investigated by Scharrer et al. (1992Scharrer et al. ( a, 1992. The present research concerns steady-state fully pressurized bearing operation.The test apparatus and facility described here is used to develop experimentally validated tools to predict steady-state hydrostatic bearing operational data (resistance torque, static load, flowrate, temperature and pressure distributions, etc.) and rotordynamic coefficients for vibration analysis. By rotordynamic coefficients, we refer to the stiffness K, damping C, and added-mass M coefficients which are used in the following linearized force-displacement model for bearingsHere, (z_, Ay) define the motion of the bearing rotor relative to its stator, and (.f,,, f+,) are the components of the fluid film reaction force acting on the rotor. Identification of the rotordynamie coefficients of Eq.(1) is a central objective of this research project.This paper describes the test-facility design requirements necessary to identify the rotordynamic coefficients and presents a sample of the resulting data for one bearing. The contents include results for a 7.62 cm (3 in.) diameter, square-recess, smooth-land, annular-fed and orifice-compensated hydrostatic journal bearing with an L/D ratio of 1 and a C,/R ratio of 0.003. The design requirements and analysis techniques for determining the steady-state operational data are detailed thoroughly by Kurtin, et a1.(1991) and will not be repeated here. Subsequ...

Analysis based on the Jeffcott model is presented to explain 1/2 speed and 1/3 speed whirling motion occurring in rotors which are subject to periodic normal-loose or normal-tight radial stiffness variations. The normal-loose stiffness variation results due to bearing-clearance effects, while normal-tight stiffness variations result from rubbing over a portion of a rotor’s orbit. The results demonstrate that 1/2 speed subharmonic motion can be explained as either a linear parametric-excitation phenomenon or as a stable nonlinear subharmonic motion. The 1/3 speed motion is shown to be possible due to the radial stiffness nonlinearity. A linear parametric-excitation analysis demonstrates that during a normal-light rubbing condition, Coulumb damping significantly widens the potential range of unstable speeds.

The basic equations are derived for compressible flow in a labyrinth seal.The flow is assumed to be completely turbulent in the circumferential direction where the friction factor is determined by the Blasius relation. Linearized zeroth and firstorder perturbation equations are developed for small motion about a centered position by an expansion in the eccentricity ratio. The zeroth-order pressure distribution is found by satisfying the leakage equation while the circumferential velocity distribution is determined by satisfying the momentum equation.The firstorder equations are solved by a separation of variables solution. Integration of the resultant pressure distribution along and around the seal defines the reaction force developed by the seal and the corresponding dynamic coefficients.The results of this analysis are compared to published test results.

Expressions are derived which define dynamic coefficients for high-pressure annular seals typical of neck-ring and interstage seals employed in multistage centrifugal pumps. Completely developed turbulent flow is assumed in both the circumferential and axial directions, and is modeled in this analysis by Hirs’ turbulent lubrication equations. Linear zeroth and first-order “short-bearing” perturbation solutions are developed by an expansion in the eccentricity ratio. The influence of inlet swirl is accounted for in the development of the circumferential flow field. Comparisons are made between the stiffness, damping, and inertia coefficients derived herein based on Hirs’ model and previously published results based on other models. Finally, numerical results are presented for interstage seals in the Space Shuttle Main Engine High Pressure Fuel Turbopump and a water pump.

An experimental test facility is used to measure the leakage and rotordynamic coefficients of teeth-on-rotor and teeth-on-stator labyrinth gas seals. The test results are presented along with the theoretically predicted values for the two seal configurations at three different radial clearances and shaft speeds to 16,000 cpm. The test results show that the theory accurately predicts the cross-coupled stiffness for both seal configurations and shows improvement in the prediction of the direct damping for the teeth-on-rotor seal. The theory fails to predict a decrease in the direct damping coefficient for an increase in the radial clearance for the teeth-on-stator seal.