Figure 5.  The Stribeck Curve for the Bearings on the Failed  Pump. bearing, the first challenge is to determine a reasonable minimum film thickness for the bearings in a  particular application. Typically, there exists a range in which the minimum film thickness may fall and the true value depends on a variety of factors. Numerically, the range for minimum film thickness lies between 7.5 and 2.5 micrometers (0.0003 and 0.0001 in). The large variation in this range is attributed to specific  machine characteristics such as surface finish of the bearing and shaft, wear couple, expected rotor motion due to various excitation forces, and lubricant rheology. In the present case, the first three characteristics are very good, the surface finish of the bearings, wear couple, and rotor motion tend toward an acceptance of low film thicknesses. However, the low lubricant viscosity and high density do  not permit the most aggressive minimum film thickness to be chosen. For this application two assumptions for minimum film thickness are used, 3.8 and 7.5 micrometers. These two levels of film thickness bound the hydrodynamic problem by providing an aggressive and a conservative estimate to  bearing performance. A Stribeck curve containing calculated performance for both front and rear journal  bearings is shown in Figure 5. Figure 6 contains results of the  bearing load capacity calculation for the rear bearing as a function of pump speed for both levels of minimum film thickness. This figure clearly demonstrates the desire to operate at the smallest possible film thickness because the hydrodynamic load capacity of the bearing roughly follows an inverse square trend with film thickness. Therefore, a welldesigned, well-built, welllubricated, smooth running machine would have over four times the load capacity of an inferior machine running in a  challenging environment. Although the analytic models for bearing performance correlate well with industry standards on low viscosity bearings, the models have not been validated with test data. As the design life of these pumps is  driven to fifteen years and beyond, it is recommended to quantitatively validate bearing performance and dynamic coefficients with component testing. Pump Rotordynamic Characteristics A rotordynamic analysis was conducted on the ISS cooling pump and the graphical results are shown in  Figure 7. Computed bearing coefficients for stiffness and damping were utilized since no individual bearing test data was available for these bearings. The composite, permanent magnet shaft was modeledwith discrete elements and the bend frequency was compared to static vibration tests to verify that total  shaft stiffness was consistent from model to test. Good agreement was achieved between test and modeling. The first two rigid body modes are shown in figure 7. A notable feature of these two modes is  that they are both angular modes. In other words there is no cylindrical mode present in this machine and,  most importantly, the machine is predisposed to angular misalignment and vibrations. Most likely this is  caused by the asymmetric journal bearing configuration where both the size and absolute clearance is different between the two bearings. This feature is not desirable for long wear life since angular  misalignment leads to edge loading on the bearings and high wear rates. For future pump designs it is  recommended to maintain cylindrical rigid body modes to reduce the high wear caused by angular contact. Additionally, qualification testing is recommended to quantitatively validate the rotordynamic models.   455