The lubricant used for the bearings was NYE 2001 Penzanne synthetic oil, Nye Lubricants, Fairhaven,  MA. The bearings were cleaned and the retainers impregnated with the test oil [2]. A thin film of oil was cast on all surfaces from immersion in dilute solution, and excess oil was removed by centrifugation at  3000 rpm. The dry bearing weight is measured and compared with the weight after lubrication. The difference represents the lubricant weight. Table 2.0.1 documents the initial lubricant weight in each test bearing.  3.0  Test Results:  PART 1 – Effect Of  Lubricant On Bearing Thermal Conductance /g3 The presence and location of lubricant affects bearing thermal properties. This relationship was explored  here by first studying a base case of a dry (no oil), static (non-moving) bearing in vacuum. We then  introduced oil into a static bearing to draw a comparison. Data was collected and an example compared the analytical predictions with that of experimental thermal conductance measurements for a small  bearing size.  3.1 Test Method All tests were conducted under steady state thermal environmental conditions, where the thermal and  mechanical boundary conditions were fixed and temperatures were allowed to equilibrate. Bearing thermal conductance measurements in section 3.2 through 3.4 were made under static conditions, 0 rpm, and dynamic measurements in section 3.5 were made at 6000 rpm. One of the challenges of the dynamic tests is that the variables, temperature and speed, are not designed to operate independently. To resolve this issue, a matrix of tests was used to perform a parametric study, where both parameters were varied.This enabled the sequential isolation of single independent variables. In this section, the test data set presented is for the average bearing temperature of 20 °C, determined by interpolation over a range of  test data. 3.2 Dry Static Bearing Bearings are generally thermally insulative. The reason for the high thermal resistance is the bearing  geometry. Two small contact areas are present, per ball, at the ball-to-inner race and ball-to-outer race interfaces generating a thermal constriction region, through which the heat must pass. A closed form  analytical model to calculate the effect of the thermal constriction was developed by Yovanovich [3].Yovanovich calculated the thermal resistance across a dry, static bearing by modeling the ball and races  as semi-infinite half-planes with the elliptical Hertzian contact area serving as the thermal contact region. The basic equations to calculate thermal resistance across each of the ball-to-race contacts are: R i = /g60i/4k 1ai + /g60i/4k 2ai inner race to ball thermal resistance (3.1) Ro = /g60/g82/4k 2ao +/g3/g60/g82/4k 1ao ball to outer race thermal resistance (3.2)  Rb= (R o+ R i) total thermal resistance across the bearing  (3.3)  where/g60is a non-dimensional geometric factor accounting for the ball contact areas and defined as: /g179 /g184/g184 /g185/g183 /g168/g168 /g169/g167/g16/g16/g32/g602/ 02/1 2 22 2 sin 12/g83 /g84/g84 /g83 nn nn ab ad (3.4) Where the subscript /g81is designated either /g76or /g82for the inner or outer race. The Hertzian contact area is  elliptical in shape and the geometry is represented by: na– major Hertzian contact axis nb– minor Hertzian contact axis 116