M. M. Khonsari
Edited by Jessica Shapiro
“Maximum Temperature for Hydrodynamic Bearings under Steady Load,” E.R. Booser, C.L. Linkinhoker, and F.D. Ryan, Lubrication Engineering, Vol. 26, pp. 226-235, 1970.
“Cool Ideas for Sleeve Bearings” (Machine Design, April 23, 2009) addressed modifications engineers can make to keep bearings cool, prolonging the lives of lubricants and equipment. Excessive heat buildup is not limited to sleeve bearings; thrust bearings can also benefit from some heat-dissipation techniques we investigated.
Similar to sleeve bearings, the performance of fluid film thrust bearings substantially deteriorates as operating temperatures rise. Oil viscosity in the bearing clearance drops drastically due to viscous heating in the tight film gaps. As a result, the oil film thins and is less able to cool the bearing. These factors, in turn, increase the likelihood temperatures will be high enough to soften bearing material and even result in bearing wipe, smearing of the bearing material along the contact surface.
Lowering thrust-bearing temperature is particularly crucial in rotating-compressor and turbine designs where designers do not know the exact net thrust on the bearings.
Designers can apply a number of cooling factors to the four common types of thrust bearings — flat land, step, tapered land, and pivoted pad. We looked at the magnitude of cooling for each technique using 5 and 10-in.-OD bearings of each type. In all cases, we assumed the bearings were running at 1,800 rpm and had a 2:1 OD:ID ratio. For more-detailed explanations of our calculations, refer to the references listed under Resources.
Under these conditions, we expect laminar oil films to form. Films start to become turbulent when the local Reynolds number in the oil film reaches 1,000 to 1,200, usually when center-line surface velocities are 10,000 to 15,000 fpm.
Process engineers typically monitor thrust-bearing mean temperature by measuring the temperature of oil draining from the unit. For critical thrust bearings, maximum temperature is often measured with a thermocouple or resistance temperature detector embedded just under the bearing surface at the 75/75 location: 75% circumferentially beyond the leading edge of a bearing segment and 75% radially outward from its inside diameter. The difference between oil-feed temperature and the reading at this 75/75 location is commonly about twice the mean temperature rise for the bearing system.
Keeping it cool
So how do designers minimize temperature rise? In thrust bearings, as in most bearing designs, the two basic ways to cool the units and cut effective running temperature are to increase oil-flow rate and reduce bearing friction.
Some friction within a bearing comes from the viscosity of the lubricant itself. Lowering lubricant viscosity can cut viscous friction and corresponding power loss. But any benefit is usually negated by a drop in oil-film thickness and a higher oil-film shear rate. Still, less-viscous lubricants can improve oil-cooler efficiency and lower friction in the oil-distribution system.
To evaluate the temperature rise of different types of thrust bearings, we selected an ISO VG-68 industrial lubricating oil with 68-centistoke viscosity at 104°F and with an oil-feed temperature of 120°F. Because the type and viscosity of most oils has relatively little effect on temperature rise, results should be representative for most conventional oils.
Increasing the volume of oil passing over each pad segment is the most-effective way to lower temperature in almost all types of thrust bearings. Designers can incorporate higher oil volume by increasing the taper over a segment, pressurizing oil feed in oil-distributing grooves, and minimizing the bearing outside diameter to keep down net surface velocity.
Finally, engineers can directly cool the oil feed or the bearing housing from the outside. Minimizing carryover of hot oil from the trailing edge of one segment into the next segment’s leading edge can be helpful. External cooling fins on the bearing housing lower overall bearing temperature, as does blowing a stream of cooling air at the housing.
When surface speeds exceed those at 3,600 rpm for 10 to 12-in. pads, parasitic power losses from churning oil in the bearing housing become significant. To avoid these losses and their heating effect, oil can be introduced directly into the leading edge of each bearing segment via leading-edge oil spray or leading-edge feed grooves. This cuts down on churning and decreases the amount of hot oil being carried from the trailing edge of one pad into the entrance edge of the next pad.
Designers may also choose materials for the bearing-support structure in part based on their thermal conductivity. For instance, copper plates, rather than the steel backing commonly used in babbitted thrust bearings, can drop the maximum temperature by 5 to 10°F.
These techniques work for all types of thrust bearings, but designers should take some type specific considerations into account as well.
Flat-lands and steps
Flat-land thrust bearings are commonly used for relatively low loads, in the range of 20 to 100 psi. Flat-land bearings’ simplicity has led to their use in electric motors and related machines and as back-up bearings that handle incidental reverse-thrust transients in a variety of applications.
Flat-land bearings’ geometry offers no distinct pumping action for entering feed oil. Rather, they rely on minor thermal expansion of the bearing material and the oil itself along the circumferential path on each bearing segment to generate load-carrying capacity. The flat thrust face limits load per unit bearing area to about 30 psi. This rises to 100 psi if the bearings have radial oil-distributing grooves.
Adding these grooves can keep bearing temperature down. Mean bearing temperature drops from 225°F with four radial grooves to 157°F when using 10 grooves. The flat lands of the bearings mean that oil-film geometry is hard to define clearly. Therefore, values calculated in the accompanying table contain some uncertainty, but they do reflect general experience with industrial machinery.
Adding a simple step to an otherwise flat bearing can help generate oil-film pressure that supports greater bearing load, sometimes up to 300 psi. For typical “square” segments where radial length L equals circumferential breadth B, the optimum step location is at 50 to 55% of B from the leading edge of the segment. Step height should be 0.7 times the trailing edge film thickness, h2.
Step designs are best used in small bearings where coining, etching, or stamping of a simple thrust washer or machine surface can form the optimum small step height. In large machinery, however, dirt accumulation and step wear make this type of bearing less effective over time. To improve cooling of a step bearing, shrouds along the sides of the inlet section (see graphic) can increase load capacity up to 70% as minimum film thickness at a given load grows by 30%.
Tapered lands and pivoted pads
For slightly larger outlet oil gaps, many bearings use four or more tapered segments. The oil wedge the tapers provide creates a pumping action that builds up oil pressure for greater load support in each bearing segment.
Oil-film thickness is defined at the inlet by h1, and at the outlet by the smaller h2. For 5 and 10-in.-OD bearings, our work indicates bearing temperature decreases with increasing tapers on individual lands. The lowest maximum bearing temperatures are usually found when the h1:h2 ratio is between 2:1 and 4:1.
Larger tapers continue to lower mean bearing temperature as more inlet oil flows over each land. But beyond the 4:1 ratio, h2 can get too small, boosting temperatures at the trailing edge. A flat land of 10 to 20% of the circumferential length, incorporated at the trailing edge of each segment, can mitigate this thinning and provide thrust support during start-up and shut down.
The values given in the accompanying table are for a fixed speed. But because h2 varies with speed, the required taper for optimum oil-pumping action and load capacity varies along with it. To accommodate a range of operating speeds, designers commonly choose pivoted-pad bearings that automatically adjust their h1:h2 ratio for steady performance at varying speeds.
A pivoted pad provides much the same performance as a fixed, tapered land with the same inclination. As with the fixed-land units, the bearing is divided into a set of “pieshaped” segments with circumferential breadth B of each sector approximately equal to radial length L. Many off-the- shelf bearings consist of six pads with an outer radius equal to twice their inner radius and with L/B = 1.
The supporting pivot is set 55 to 58% radially outward from the inner radius of a pad and 50% circumferentially along from the leading edge. These locations help keep the pad from tilting radially and let it operate in either direction of rotation. They also provide a useful h1:h2 of about 2:1 for widely varying loads, rotational speeds, and oil viscosities.
In many bearings, the pads deflect into an umbrella shape under heat and load. This shape creates a gap similar in magnitude to the minimum film thickness, providing a useful oil-film wedge shape even when the pivot is centrally located.
Pivots shifted to positions 70 to 75% circumferentially beyond the leading edge provide better cooling performance, however. The accompanying table shows the relationship between pivot location and bearing temperature.
Mean-temperature rise is essentially independent of rotational speed, oil viscosity, and overall bearing size. So designers can approximate the bearing thrust load by applying a simple multiplier to the difference between the mean bearing temperature and the oil inlet temperature, as in:
P = k × ( Tmean)
Values of the multiplier k are shown in the accompanying table for different bearing types. While the values are quite general for various speeds, bearing diameters, and oil viscosities, the table specifically reflects the case of VG-68 mineral oil fed at 120°F for six-pad bearings with a 10-in. OD and 5-in. ID where L/B = 1.
For example, designers could calculate the load leading to a 50°F temperature rise in a 100-in.2-pad bearing with 50% pivots as follows:
P = k × ( Tmean) = 11 × (50) = 550 psi or 55,000-lb axial force.
Examining k shows that the dimensionless multiplier of a flat thrust-surface bearing increases from 1 to 11 when a 2:1 taper is introduced. With centrally pivoted pads, this same value is available for both directions of rotation and throughout a range of variable speeds. To reach the optimum k value of 27, pivoted pads, where the pivot is offset to 60 to 70%, offer this high load factor throughout a range of variable speeds.