Test rigs that replicate the conditions for thrust collars (TCs) used in an integrally geared compressor (IGC) are scarce. The test rig described here is based on a typical IGC and is the first rig specifically designed to measure the dynamic reaction force coefficients of the lubricated area of the TC. The test rig uses low-speed and high-speed shafts with independently controlled speed and a pneumatically pressurized thrust disk to apply an axial load ̅ to create the hydrodynamic wedge that balances the imposed axial load. The speed ratio between the low-speed shaft (LSS) and the pinion shaft is 11.67.The geometry of the shafts matches that of a typical IGC. Tests were conducted at pinion speeds of 5, 7.5, and 10 krpm and ̅ = 200, 300, and 400 N. The resulting range of applied pressures is smaller than those arising in practice.The author conducts static tests by applying an incrementally-increasing ̅ on the pinion shaft and measuring the relative displacement between the BG and the TC (Δ ̅ ). One test is conducted at each predetermined spin speed. Run-out on the TC as well as the BW obscures the data. Averaging works well to eliminate the effects of run-out.The author uses the averaged ̅ and Δ ̅ values to create a static, load/ relativedisplacement curve and the slope is the measured static stiffness coefficient ( ̅ ). The axial stiffness coefficient results are compared to predictions from a code based on a 2016 model due to Cable et. al. Their dynamic reaction-force model is = − Δ − Δ̇ iii where is the reaction force of the TC, and is the axial damping coefficient. The trends and the magnitudes of the measured ̅ values and the predicted values from San Andres code for agree very well, especially for the 5 krpm test case. The author then conducts dynamic tests involving an applied impulse load to the TC shaft. One hundred impulses are conducted at each spin speed ( ), ̅ test condition for averaging purposes. A one degree of freedom damped motion model uses Δ ( ) measurements to determine the damped natural frequency ( ) and damping factor ( ) for each test point. The thrust collar mass and the measured were then used to calculate and . The values obtained in this fashion were consistently (and markedly) smaller than the static ̅ values. Based on the results, the author uses the following model = − Δ − Δ̇− Δ̈ that includes the virtual-mass coefficient ( ). The Cable et al. model was based on the Reynolds equation and accordingly did not produce a virtual-mass term.The term is calculated for each test point using ̅ , , and . increases as a function and ̅ . It ranges from 0 to 19.5 kg; the mass of the pinion shaft is 12.8 kg.Both predictions and measurements show an increase in with increasing ̅ . The test rig produced damping coefficients that increased for increasing , while the predicted values decreased. The magnitude of was lower than the predicted damping by a factor of 2 -10.iv ACKNOWLEDGEMENTS

Sliding seals play a critical role in the dynamic sealing of a wide-variety of machinery applications. There are many type of sliding seals, such as segmented carbon ring, O-ring, or bellows; and are used as sealing elements in hydraulic rod and piston seals, as well as secondary sealing elements in mechanical and dry-gas seals. Motion between the dynamic and static parts of the machine is most often lubricated by the process fluid, and therefore has leakage. This paper presents a new test rig capable of measuring leakage and friction force of annular sliding seals for a range of sealing pressures (0–24.8 MPa), temperatures (20–700 C), gasses, and motion. The rig is capable of large linear motions and high frequency dynamic motion. The large linear motion replicates piston movement or shaft thermal growth and the dynamic excitation matches the typical vibration response from a spinning shaft. Currently, there is little available information in the literature on the leakage performance of dynamic O-rings at high pressures, especially for gas sealing. The paper presents experimental results from two different material, 150 mm diameter, O-rings at sealing pressures up to 70 bar (1,000 psi) in CO2. The rings were tested for large ranges of motion, up to 4 mm. Two different durometer FKM O-rings (70 and 90) were compared to a 70 durometer BUNA O-ring. The softer rings exhibited superior leakage performance, and similar friction forces. The BUNA O-ring performed slightly better at sealing (10–20%) then the similar hardness Viton ring.

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Windage is the effect of aerodynamic drag on the surfaces of a rotating system due to fluid shear effects. The fluid-friction losses that occur on the rotor of rotating machines often constitute a non-negligible drag on the system that must be estimated for proper sizing of the driving or driven element. This is especially true in high-pressure environments, such as hermetic compressors and turbines. Fluid-friction loss modeling is based on the size and rotation speed of the shaft, the density of the fluid, and an empirically-determined drag coefficient. The drag coefficient is generally a function of the Reynolds number but may also be dependent on the Taylor number. Several papers have provided empirical predictions for drag coefficients based on the Reynolds and Taylor numbers of the fluid, but other factors such as rotor shapes, assemblies, and surrounding fluid conditions can also affect the drag coefficient. There are two main geometries for a rotor: a face parallel to the axis of rotation, and a face that is perpendicular. The gap between the rotating component and the stationary housing also plays an important role in the drag coefficient. This review summarizes and compares these findings in a way that makes it easy for the reader to predict the total windage losses on a system for any rotor shape, speed, or operating condition. A quick reference table is presented in the conclusions section.