Stanley I. Pinel

Pinel Engineering, Fullerton, California

Hans R. Signer

Signer Technical Services, Fullerton, California

Erwin V. Zaretsky

Glenn Research Center, Cleveland, Ohio

Comparison Between Oil-Mist and Oil-Jet

Lubrication of High-Speed, Small-Bore,

Angular-Contact Ball Bearings

NASA/TM—2001-210462

August 2001
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NASA/TM—2001-210462

August 2001

National Aeronautics and

Space Administration

Glenn Research Center

Prepared for the

56th Annual Meeting

sponsored by The Society of Tribologists and Lubrication Engineers

Orlando, Florida, May 20–24, 2001

Stanley I. Pinel

Pinel Engineering, Fullerton, California

Hans R. Signer

Signer Technical Services, Fullerton, California

Erwin V. Zaretsky

Glenn Research Center, Cleveland, Ohio

Comparison Between Oil-Mist and Oil-Jet

Lubrication of High-Speed, Small-Bore,

Angular-Contact Ball Bearings
Available from

NASA Center for Aerospace Information

7121 Standard Drive

Hanover, MD 21076

National Technical Information Service

5285 Port Royal Road

Springfield, VA 22100

Available electronically at http://gltrs.grc.nasa.gov/GLTRS
NASA/TM—2001-210462 1

Parametric tests were conducted with an optimized 35-mmbore-angular-contact ball bearing on a high-speed, hightemperature bearing tester. Results from both air-oil mist

lubrication and oil-jet lubrication systems used to lubricate

the bearing were compared to speeds of 2.5×106 DN. The

maximum obtainable speed with air-oil mist lubrication is

2.5×106 DN. Lower bearing temperatures and higher power

losses are obtained with oil-jet lubrication than with air-oil

mist lubrication. Bearing power loss is a direct function of oil

flow to the bearing and independent of oil delivery system.

For a given oil-flow rate, bearing temperature and power loss

increase with increases in speed. Bearing life is an inverse

function of temperature, the difference in temperature between the individual bearing ring components, and the resultant elastohydrodynamic (EHD) film thicknesses. Bearing life

is independent of the oil delivery system except as it affects

temperature. Cage slip increased with increases in speed.

Cage slip as high as 7 percent was measured and was

generally higher with air-oil mist lubrication than with oil-jet

lubrication.

Comparison Between Oil-Mist and Oil-Jet Lubrication

of High-Speed, Small-Bore, Angular-Contact Ball Bearings

Stanley I. Pinel

Pinel Engineering

Fullerton, California

Hans R. Signer

Signer Technical Services

Fullerton, California

Erwin V. Zaretsky

National Aeronautics and Space Administration

Glenn Research Center

Cleveland, Ohio 44135

INTRODUCTION

Grease and simple splash lubrication are the lubrication

means of preference for most low-speed machinery applications where the bearing speeds are less than 200 000 DN.

(DN is a rolling-element bearing speed parameter where

D represents the bearing bore in millimeters and N the bearing

inner-ring speed in rpm.) For aircraft engines, turbomachinery,

and drive trains, jet lubrication is used both to lubricate

bearings and gears and to control their temperature by removing generated heat. In jet lubrication the placement and

number of nozzles, the jet velocity, the lubricant flow rates,

and the removal of lubricant from the bearing and immediate

vicinity are all important for satisfactory operation, Zaretsky

(1).

A more effective and efficient means of lubricating rollingelement bearings is under-race lubrication so that increasing

the oil flow decreases the ring temperature. Conventional jet

lubrication fails to adequately cool and lubricate the innerrace contact because lubricant is thrown centrifugally

outward. While increasing flow rate results in carrying away

more heat, it also adds to the heat generated from oil churning,

Zaretsky (1).

Ring cooling by oil flow at the exterior ring surfaces has

been used for many years for high-speed turbomachinery

bearings. To achieve speeds of 3 million DN with angularcontact ball bearings and cylindrical roller bearings, Zaretsky

et al. (2) and Signer et al. (3) proposed the concept of “bearing

thermal management” as the proper technological approach

to high-speed bearing operation. Bearings lubricated through

the inner ring passages were also provided with cooling of the

external outer ring surfaces. The basis of this concept was the

recognition that total and flexible thermal control over the

bearing components was essential to achieving a reliable

high-speed, highly loaded bearing, Zaretsky (1).
NASA/TM—2001-210462 2

Both the conventional jet and the under-ring lubrication

systems recirculate the oil used as the lubricant. There are

rotating machinery applications in which a nonrecirculating

lubrication system, which operates with minimum possible oil

flow, could be employed advantageously. Weight and cost

reduction of the system might be realized. One such lubricating system that has been used commercially for decades is airoil mist lubrication. Oil-mist lubrication systems have been

used extensively for industrial applications, high-speed machine spindles, and in limited aerospace applications. Bloch

and Shamim (4) describe these oil-mist lubrication systems

together with performance data.

Early work by Schuller and Anderson (5,6) studied and

determined the effects of minimum oil flow with air-oil mist

lubricated, thrust-loaded, 75-mm angular-contact ball bearings to 1.2×106 DN under varying loads, oil-flow rates and

temperature. They found that the bearings operated satisfactory at very low oil flows with air-oil mist lubrication using a

synthetic diester lubricant (MIL–L–7808). Minimum oil-flow

requirements increased with increases in load and speed. They

developed for an air-oil mist lubrication system an empirical

formula relating minimum oil-flow requirements as a function

of load and speed for successful ball bearing operation.

Research performed by Morrison et al. (7) with 45-mm

angular-contact ball bearings run to speeds of 135 000 DN run

with sump lubrication and air-oil mist found that the air-oil

mist lubrication system resulted in lower bearing temperatures on the order of 12 to 15 °C (22 to 27 °F). Shamim and

Kettleborough (4,8) also compared an air-oil mist lubrication

system with a conventional sump lubrication system for

60-mm-bore-angular-contact ball bearings run to speeds

of 144 000 DN. Their results were similar to those of

Morrison et al. (7) where the air-oil mist lubricated bearing ran

10 °C (18 °F) cooler than the sump lubricated bearing and at

a lower power loss.

Rosenlieb (9) performed once-through, air-oil mist lubrication studies with 125-mm-bore-angular-contact ball bearings

made from AISI M–50 steel. The lubricants used in his study

were two synthetic paraffinic hydrocarbon fluids and a Type

II advanced ester fluid. He demonstrated the feasibility of airoil mist lubrication to speeds of 2.5×106 DN at bearing

temperatures of 216 °C (420 °F) for periods of time up to

35 hr at an oil-flow rate as low as 3.7×10–3 /min. Increasing

the air-oil mist temperature resulted in an equal increase in

bearing temperature.

In a follow-on study, Rosenlieb (10) demonstrated that an

air-oil mist system along with auxiliary air cooling could

provide short-term operation of a 46-mm-bore-angularcontact ball bearing to speeds up to 3×106 DN and temperatures up to 264 °C (505 °F) with flow rates as low as

0.9×10–3 /min. However, at speeds greater than

2.5×106 DN bearing cage instability and wear occurred. The

cage instability occurred only with oil-mist lubrication and

not with oil-jet lubrication. He attributed the cage instability

to the reduced oil damping provided by the air-oil mist. The

lubricants used in this study were two different Type I esters

meeting the MIL–L–7808 specification, a Type II ester meeting the MIL–L–23699 specification, and a polyphenylether.

Rosenlieb (10) also ran air-oil mist bearing lubrication

endurance tests at 2×106 and 2.5×106 DN for 100 and 50 hr,

respectively.

Jeng and Huang (11) compared the temperature rise of

hybrid ceramic and steel 50-mm-bore-angular-contact ball

bearings run to 750 000 DN. For their tests they varied load,

speed, airflow rate, and oil quality. Their test results compared

well with those of Schuller and Anderson (5,6). Increasing the

airflow rate for a constant lubricant flow marginally reduced

bearing temperature. Bearing temperature increased with speed

and to a lesser extent with load and decreased lubricant flow

rate. The hybrid ceramic bearing ran at lower temperatures

than the steel bearing by 4 to 8 °C (7 to 14 °F). However, as the

lubricant flow rate is increased, the temperature difference

becomes inconsequential. This can be attributed to the volume

of oil entrapped in the bearing beyond that necessary for

elastohydrodynamic lubrication (EHD). Initially, the smaller

Hertzian contact ellipse with the hybrid ceramic bearing

results in lower frictional losses. Churning of the oil at the

higher flow rates accounts for the major frictional loss in the

bearing resulting in both bearing types having similar temperature values.

In Pinel et al. (12) we reported the results of our research

with high-speed, small-bore, angular-contact ball bearings

run to 2.5 million DN with both oil jet and under inner-ring

lubrication. The results of this research optimized the design

of oil-jet lubricated 35-mm-bore-angular-contact ball bearings to a single-outer-ring, land-guided cage with a 24o contact

angle with a relieved inner ring and a partially relieved outer

ring. This design resulted in lower temperature and power

loss.

The objectives of the work reported herein are to (1) extend

our work with the 35-mm-bore-angular-contact ball bearings

to include the effect of air-oil mist lubrication; (2) define and

extend ball bearing operating limits with air-oil mist lubrication; and (3) compare air-oil mist lubricated angular-contact

ball bearing operation with oil-jet lubricated bearings to

2.5×106 DN.

APPARATUS, SPECIMENS, AND PROCEDURE

High-Speed Bearing Tester

A schematic drawing of the air-turbine-driven test machine

is shown in Fig. 1(a) and was initially described by Pinel and

Signer (13). The test rig consisted of a horizontally mounted

shaft supported by two preloaded, angular-contact ball bearings. The test bearing was overhung and mounted in a separate

housing (Fig. 1(b)) that incorporated the hardware for

lubrication, oil removal, thrust load application, and instru-
NASA/TM—2001-210462 3

Figure 1.—High-speed, small-bore-bearing tester. (a) Schematic. (b) Test bearing installation.

Inner-ring

temperature

telemetry

system

Thrust-load

sensor

Shaft speed

pickup Outer-ring

cooling oil

Outer-ring

thermocouple

Lubricant oil

supply to support

bearings

Test bearing

Thrust

load

Discharge-oil

thermocouple

Drain Drain

Support

bearing

Support

bearing

Turbine

nozzle

Air-driven

turbine

Inner-ring

thermocouple

(a)

(b)

Air/oil mist supply line

Test bearing

Annular reclassifier

Thrust load

Plug
NASA/TM—2001-210462 4

mentation for cage speed measurement. The test-bearing torque

was measured with strain gages located near the end of an arm

that prevented the housing from rotating. Thrust force was

applied through the combination of a thrust needle bearing and

a small roller support bearing to minimize test-housing restraint during torque measurements.

Test Bearing

Based upon the work reported in Pinel et al. (12), a 35-mmbore-angular-contact ball bearing with relieved inner and

partially relieved outer rings, a 24° unmounted contact angle,

and a single-outer-ring, land-guided cage (Fig. 2) was used for

the tests reported herein. The specifications for the test bearings are summarized in Table 1. The bearings were an ABEC

7 grade and each contained 16 balls with a nominal 7.14-mm

(0.281-in.) diameter. The inner and outer rings and the balls

were manufactured from consumable-electrode, vacuummelted, AISI M–50 steel. The surface finishes of the races were

0.125 μm (5 μin.) rms or better. The surface finish of the balls

was 0.025 μm (1 μin.). The bearing surface composite finish

was 0.130 μm (5.1 μin.). The nominal hardness of the balls and

rings was Rockwell C62 at room temperature.

The cage was manufactured from AISI 4340 steel (AMS

6415) heat-treated to Rockwell C 28 to 36 hardness. It was

completely coated with a 0.02- to 0.04-mm (0.0008- to

0.0015-in.) thickness of silver plate (AMS 2412). The cage

balance was within 0.50 g-cm (7×10–3 oz-in.).

Lubricant

The lubricant used for the parametric studies was a

neopentylpolyol (tetra) ester. This Type II oil is qualified to the

MIL–L–23699 specifications and also to the internal oil specifications of most major aircraft engine manufacturers. The

major properties of the lubricant are presented in Table 2.

Thrust Cage

Oil

jet

Inner

ring

Outer ring

Thrust

Figure 2.—35-mm-bore, angular-contact ball bearing

 with relieved inner ring (design 4). (a) Single outer-

 ring, land-guided cage. (b) Bearing assembly.

(a)

(b)

Disassembled Assembled

Radial

lubricant

supply hole

Axial oil

groove

TABLE 1.—TEST BEARING SPECIFICATIONS

Bearing

Bore, mm (in.)...................................................................35 (1.3780)

Outside diameter, mm (in.)...............................................62 (2.4409)

Width, mm (in.).................................................................14 (0.5512)

Grade......................................................................................ABEC 7

Surface composite finish, μm (μin.)..................................0.130 (5.1)

Cage

Diametral land clearance, mm (in.)

Outer-land-guided .....................................................0.406 (0.016)

Inner-land-guided......................................................0.254 (0.010)

Diametral ball pocket clearance, mm (in.)...................0.660 (0.026)

Material..............................AISI 4340 per AMS 6415 (silver plated)

Hardness Rockwell C..............................................................32 to 38

Balls

Number.............................................................................................16

Size (diam), mm (in.).......................................................7.14 (0.281)

Grade................................................................................................10

Material........................................CEVM AISI M–50 per AMS 6490

Hardness, Rockwell C (nominal)..........................................62 (min.)

Surface finish, μm (μin.) rms...............................................0.025 (1)

Race

Conformity, percent

Inner.............................................................................................54

Outer............................................................................................52

Surface finish, μm (μin.), rms..............................................0.125 (5)

Nominal contact angle, deg.............................................................24

Internal diametral clearance (IDC), μm (μin.)

Relieved inner race......................................................0.074 (0.0029)

TABLE 2.—PROPERTIES OF NEOPENTYLPOLYOL

(TETRA) ESTER LUBRICANT

Additives

Corrosion and oxidation inhibitors and antiwear and antifoam

additives

Kinematic viscosity, cS, at—

38 °C (100 °F)............................................................................... 28.5

99 °C (210 °F)............................................................................... 5.22

204 °C (400 °F)............................................................................. 1.31

Flashpoint, °C (°F)....................................................................260 (500)

Autogenous ignition temperature, °C (°F) ...............................421 (800)

Pour point, °C (°F) ....................................................................–59 (–75)

Volatility (6.5 hr at 204 °C (400 °F)), wt%.........................................3.2

Specific heat at 99 °C (210 °F), J/kg·°C (Btu/lb·°F)............ 2140 (0.493)

Thermal conductivity at 204 °C (400°F),

J/m·sec·°C (Btu/hr·ft·°F)...................................................0.13 (0.075)

Specific gravity at 99 °C (210 °F) .................................................. 0.931

Specification ....................................................................MIL–L–23699
NASA/TM—2001-210462 5

Test Procedure

After the test machine had been warmed by recirculating

heated oil through the support bearings, the torque-measuring

system was calibrated and a 667-N (150-lbf) thrust load was

applied to the test bearing. The shaft speed was then slowly

brought up to a nominal speed of 28 000 rpm. When the bearing

and test machine temperatures stabilized (after 20 to 25 min),

the oil-inlet temperature and lubricant flow rate were set and

the speed was increased to the desired value.

A nominal thrust load of 667 N (150 lbf) was used for these

tests. For each condition, tests were run at nominal speeds of

30 000, 50 000, and 65 000 rpm. A test series was run by

starting at the lowest nominal speed, 30 000 rpm, and progressing through 50 000, and 65 000 rpm before changing the

lubricant flow rate. During testing, if it became apparent that

the conditions would result in predictable distress of the test

bearing or test rig or that the bearing temperature would exceed

a predetermined temperature, the test point was aborted.

Limiting ring temperature was set at 218 °C (425 °F) for the

recirculating oil-jet lubrication. A higher ring temperature

limit of 260 °C (500 °F) was chosen for air-oil mist lubrication,

as some degradation of the lubricant was permissible for a

“once through” lubrication system.

The shaft speed (inner-ring speed) was measured with a

magnetic probe. The ball-pass frequency (cage speed) was

measured with a semiconductor strain gage mounted in a cavity

of the housing and was displayed on a spectrum analyzer.

Two thermocouples were assembled in the shaft so that the

centrifugal force would push them against the test-bearing

inner ring. Temperature readings were transmitted with a

rotating telemetry system mounted on an auxiliary shaft at the

air-turbine end of the test machine. Outer-ring temperatures

were obtained by two thermocouples located 180° from each

other.

Lubrication System

Air-Oil Mist—A schematic diagram of the air-oil mist lubrication system used for these tests is shown in Fig. 3. The airoil mist system used was identical to that described by Schuller

and Anderson (5). By flowing pressurized air into a venturi, the

pressure in the venturi is reduced below that in the reservoir

and oil is siphoned from the reservoir into the venturi. The oil

in the venturi is atomized into a fine mist and carried into an

annular reclassifier and into the bearing. Figure 1(b) shows the

annular reclassifier and a cross section of the test bearing

installation. The drain on the reclassifier side is plugged so as

to force the air-oil mist through the reclassifier tubes and

through the bearing.

The oil flow to the test bearing was regulated by air pressure

to the lubrication tank and selection of the size of the capillary.

The oil temperature was controlled by regulating the oil tank

wall temperature. The average oil-in temperature was 49 °C

(120 °F) at the venturi entrance.

Figure 3.—Air-oil mist lubrication system.

Regulator

Hot plate heater

Inlet to

reclassifier

Test housing

Pressure gauge

T/C location

Capillary

tube

Supply tank

Flow meter

Filter

Separator

Shop air
NASA/TM—2001-210462 6

Oil Jet—For jet lubrication, the test bearing was lubricated

by two jets on the nonloaded side of the inner ring. The jet

outlets, located approximately 3 mm (0.12 in.) from the face

of the bearing, were aimed at the inner raceway. In separate

tests not reported herein, it was determined that a 20-m/sec

(66-ft/sec) jet velocity provided the most efficient test bearing

lubrication so this velocity was used in all the tests reported.

The oil-in temperature measured just before the oil jet was

121 °C (250 °F) (12).

RESULTS AND DISCUSSION

Effect of Lubricant Flow Rate

Our experience with oil-jet lubrication has shown that the

measured bearing temperature after the bearing reaches thermal equilibrium is a function of the bearing load, speed, oil-in

temperature, oil-flow rate and volume of oil entrapped in the

bearing. At thermal equilibrium, the bearing race temperatures

without inner and/or outer ring cooling will always be higher

than the oil-in temperature. As the oil-flow rate is increased,

the bulk bearing temperature should approach the oil-in temperature. That is, the ring temperature minus the oil-in temperature should diminish up to the point where the heat

generated due to power loss from increasing oil churning

exceeds the convective cooling capacity of the oil. All oil-flow

rates reported herein were kept well below this critical point.

In our experiments, we used two different oil-in temperatures, 49 °C (120 °F) and 121 °C (250 °F) for air-oil mist and

oil-jet lubrication, respectively. These temperatures were

selected as perhaps being representative of what may be used

in high-speed turbomachinery applications. Thus, in our tests

the temperature of the oil-jet lubricated bearing can never be

less than 121 °C (250 °F). Similar experience or data did not

exist for air-oil mist lubrication.

A comparison of the experimental results is shown in Fig. 4

for the test bearing for both the air-oil mist lubrication and oiljet lubrication. The measured inner- and outer-ring temperatures decreased as lubricant flow rate increased for both

lubrication methods. Because the heat transfer to the lubricant

is a function of the mass flow rate of the oil to bearing, the

resultant bearing temperature at constant operating conditions

is a function of the lubricant flow rate and the oil-in temperature independent of the oil delivery system.

Referring to Fig. 4(a), because of the lower flow rate to the

bearing at 65 000 rpm with air-oil mist lubrication, the bearing

ring temperatures are higher than those with jet lubrication

even though the oil-in temperature with air-oil mist is 72 °C

(130 °F) lower. These temperatures exceed our temperature

limitation for oil-jet lubrication of 218 °C (425 °F).

Bearing temperatures of approximately 190 °C (374 °F) at

50 000 rpm with air-oil mist flow rates between 0.02 and

0.04 /min are equivalent to those temperatures with an oil-jet

flow rate of 0.08 /min. At 0.04 /min, the bearing temperature

at a speed of 30 000 rpm is approximately the same as with that

obtained with jet lubrication at 2 /min.

Referring to Fig. 4(b), the difference between ring temperature and oil-in temperature is shown as a function of lubricant

flow rate. These are the same data as those of Fig. 4(a) but

independent of oil-in temperature. As previously discussed,

bearing temperature is an inverse function of lubricant flow

rate and independent of the lubricant delivery system. It can be

reasonably concluded that a specified bearing temperature can

be maintained with less oil flow using air-oil mist providing the

oil-in temperature is kept low. However, at the higher speed of

65 000 rpm (2.3×106 DN), the air-oil mist lubrication system

becomes marginally effective to maintain bearing temperature. This necessitates the need to use either oil-jet lubrication

or under-ring lubrication (Pinel et al. (12)).

For air-oil mist lubrication, the inner-ring temperature is higher

than that of the outer ring. This difference in temperature between

the inner and outer rings was nominally less than 12 °C (22 °F).

Higher heat generation would be expected at the ball inner-race

Hertzian contact because of the contact kinematics.

For oil-jet lubrication, the outer-ring temperature is higher

than the inner-ring temperature. The same relation was

reported in Pinel et al. (12) for under-ring lubrication. At a flow

rate of 0.09 /min at 50 000 rpm, the outer-ring temperature is

less than 18 °C (32 °F) higher than that of the inner ring. As

lubricant flow rate is increased the temperature decreases and

the inner- and outer-ring temperatures become nearly equal as

shown in Fig. 4(a).

The lubricant jets were aimed directly at the bearing inner

raceway surface. The lower inner-ring temperatures for the jet

and the inner-ring lubrication methods may be explained by

more effective convective cooling.

Referring to Fig. 5, bearing power loss as a function of oilflow rate is shown for both air-oil mist and oil-jet lubrication.

The power loss data is based on mechanical measurements of

bearing torque. Because of low oil flow with the air-oil mist

lubrication system, it was not possible to thermally determine

power loss. We reported in Pinel et al. (12) that because of the

differences in measurement techniques, the thermal measurements were consistently 0.2 kW (0.3 hp) higher than those

determined mechanically with oil-jet lubrication. These data

show that the power loss is a function of oil flow to the bearing

and, for these tests, independent of the delivery system.

Because a smaller oil flow is delivered with the air-oil mist

system, the power loss is lower than that obtained with oil-jet

lubrication.
NASA/TM—2001-210462 7

Figure 4.—Effect of lubricant flow rate on bearing

 temperature for 35–mm bore angular-contact ball

 bearing lubricated by air-oil mist and oil jet. Thrust

 load, 667 N (150 lbf); nominal contact angle, 24°;

 cage, single outer-land guided; oil-in temperature,

 49 °C (120 °F) for air-oil mist, 121 °C (250 °F) for oil

 jet. (a) Measured ring temperatures. (b) Difference

 between ring temperature and oil-in temperature.

101 100 10–1 10–2 10–3

(b) Lubricant flow rate, /min Ring temperature minus oil-in temperature, 

T, °C

120

200

220

100

80

60

40

20

0

180

160

140

(a)

101 100 10–1 10–2 10–3

Lubricant flow rate, /min

Ring temperature, °C

160

240

260

140

120

220

200

180

Temperature limitation

for air-oil mist

Temperature limitation

for oil-jet

Shaft speed, rpm

30 000

50 000

65 000

Oil-jet lubrication

Air-oil mist lubrication

Open symbols denote

 inner ring

Solid symbols denote

 outer ring

101 100 10–1 10–2 10–3

Lubricant flow rate, /min

1.6

Bearing power loss, kW 0.4

1.2

1.4

0.2

0.0

1.0

0.8

0.6

30 000

50 000

65 000

Shaft

speed,

rpm

Open symbols denote

 air-oil mist lubrication

Solid symbols denote

 oil-jet lubrication

Figure 5.—Effect of lubricant flow rate on bearing

 power loss for 35–mm bore angular-contact ball

 bearing lubricated by air-oil mist and oil jet.

 Thrust load, 667 N (150 lbf); nominal contact

 angle, 24°; cage, single outer-land guided; oil-

 in temperature, 49 °C (120 °F) for air-oil mist,

 121 °C (250 °F) for oil jet.

Air-oil mist

Oil jet
NASA/TM—2001-210462 8

Effect of Speed

Figure 6 shows the effect of speed on bearing temperature

and power loss. For a given oil-flow rate, bearing temperature

and power loss increase with increases in speed, independent

of the lubrication delivery system. Referring to Fig. 4(a), airoil mist lubrication results in high-bearing temperature

because of a lower volume of oil delivered to the bearing than

with oil jet. However, at 50 000 rpm, the lower power loss is

approximately a third of that obtained for oil-jet lubrication

and, from Pinel et al. (12), for under-ring lubrication. From the

data reported herein and those of Rosenlieb (9,10), it can be

reasonably concluded that where the bearing temperature rise

is solely caused from heat internally generated from within the

bearing cavity, the maximum obtainable speed with air-oil

mist lubrication is 2.5×106 DN.

Minimum Oil Flow

Referring to Fig. 4(a), for air-oil mist lubrication, at

2.3×106 DN (65 000 rpm), the minimum oil flow necessary to

maintain a temperature of 260 oC (500 oF) was approximately

0.02 /min. No attempt was made in these tests to determine at

the other speeds a minimum or threshold flow rate below which

the bearing temperature would exceed a temperature of 260 oC

(500 oF).

Schuller and Anderson (5,6) developed an empirical formula which we reconstituted relating minimum oil-flow requirements as a function of load and speed for successful ball

bearing operation for an air-oil mist lubrication system where

V ≈ k[P4.22×101.18(S – 8)] (1)

Where V oil-flow rate, /min

k proportionality factor

P thrust load, N

S DN value × 10–5

Equation (1) was calibrated by us at 2.3×106 DN

(65 000 rpm) and a flow rate (V) of 0.038 /min with the

35-mm-bore-angular-contact ball bearings reported herein

whereby k = 9.16×10–32. A resultant minimum oil-flow rate at

of 17.4×10–18 /min at 1 million DN is predicted using Eq. (1).

This predicted oil-flow rate is significantly lower than the rate

we used and experimentally determined by Schuller and Anderson (5,6) under comparable conditions. The relation in the

brackets relating load, P, and speed, S, appears to us to be

specific to the data of Schuller and Anderson (5) and does not

appear to be universally applicable for predicting minimum oil

flow with air-oil mist lubrication.

Air-oil mist lubrication

Oil-jet lubrication

1.910–3

3.810–3

1910–3

3810–3

9510–3

31010–3

76010–3

189010–3

Lubricant flow rate,

/min

2 4 3

(a)

5 67104

Speed, rpm

250

Outer-ring temperature minus

oil-in temperature, 

T, °C

100

150

200

50

0

Figure 6.—Effect of bearing speed on bearing

 temperature and power loss for 35–mm bore

 angular-contact ball bearing lubricated by air-oil

 mist and oil jet. Thrust load, 667 N (150 lbf); nominal

 contact, 24°; cage, single outer-land guided; oil-in

 temperature, 49 °C (120 °F) for air-oil mist, 121 °C

 (250 °F) for oil jet. (a) difference between outer-ring

 temperature and oil-in temperature. (b) Bearing

 power loss.

2 4 3

(b)

5 67104

Speed, rpm

1.6

Bearing power loss, kW

0.4

1.2

1.4

0.2

0.0

1.0

0.8

0.6
NASA/TM—2001-210462 9

Effect on Cage Speed

The cage (separator) epicyclic speed ratio for the angularcontact ball bearings changes with speed, load, and temperature. The difference between the cage epicyclic speed and the

measured cage speed is indicative of ball slip as follows:

Percent Cage Slip = [1 – (Measured Cage Speed ÷

Calculated Epicyclic Cage Speed)] × 100 (2)

Cage slip is a function of the traction in the ball-race contact

and the resistance to rolling within the bearing. Experience has

shown that both ball and roller bearings can tolerate reasonable

amounts of rolling element to raceway slip provided that a

sufficient EHD film is maintained in the ball-race contact.

However, under marginal lubrication conditions, such slip can

result in bearing damage and failure.

For the bearing used in this study, under a thrust load of

667 N (150 lbs) and speeds of 30 000, 50 000, and 65 000 rpm,

the epicyclic cage speeds were calculated to be 13 600, 24 000,

and 33 100 rpm, respectively. Based upon measured cage

speed, the percentage of cage slip was calculated and plotted as

a function of bearing speed for air-oil mist lubrication and

oil-jet lubrication in Fig. 7. Cage slip increased with increases

in speed. Values of cage slip as high as 7 percent were

measured. There was no damage to the bearing surfaces as the

result of cage slip. Contrary to what we expected, cage slip was

generally higher with air-oil mist lubrication than with oil-jet

lubrication.

Effect on Bearing Fatigue Life

We calculated the fatigue life as a function of speed based on

measured ring temperature with oil-jet lubrication and underring lubrication for the 35-mm-bore-angular-contact ball bearing used in this study. We reported and discussed the results in

Pinel et al. (12). As is commonly expected, life decreased both

in total revolutions to failure and in absolute hours as a function

of speed.

Shamim and Kettleborough (4,8) reported on endurance

tests with air-oil mist lubrication and industrial sump lubrication systems. They reported that the air-oil mist lubricated

bearings ran approximately 10 oC (18 oF) cooler than those

with sump lubrication and with a 25-percent reduction in

power loss. They ran endurance tests with 60-mm-bore-angular-contact ball bearings at 2400 rpm (144 000 DN) and a thrust

load of 18 900 N (4250 lbs) using an air-oil mist lubrication

system and a synthetic ISO Grade 68 oil. They compared their

results with those calculated for the same bearing using an oil

sump lubrication system with a mineral oil base lubricant.

They concluded that longer life is obtained with air-oil mist

lubrication with the synthetic lubricant than with the same

bearings operated with the mineral oil and a sump lubrication

system.

For a given bearing load and speed, the bearing life is

affected by the internal diametral clearance at operating temperature and the EHD film thickness. The method of lubrication used can and apparently does affect these parameters.

Thus, while the Shamim and Kettleborough (4,8) conclusion is

correct for their experimental conditions and assumptions

made for their calculations, it may not always be correct for

other bearing operating conditions.

Air-oil mist lubrication

Oil-jet lubrication

1.910–3

3.810–3

1910–3

3810–3

9510–3

31010–3

76010–3

189010–3

Oil flow,

/min

Figure 7.—Effect of lubrication means and speed on

 cage slip for 35–mm bore, angular-contact ball

 bearing lubricated by air-oil mist and oil jet. Thrust

 load, 667 N (150 lbf); nominal contact angle, 24°;

 cage, single outer-land guided; oil-in temperature,

 49 °C (120 °F) for air-oil mist, 121 °C (250 °F) for

 oil jet.

3 5 4 6 78104

Speed, rpm

7.00

Cage slip, percent

2.00

3.00

6.00

1.00

4.00

5.00
NASA/TM—2001-210462 10

Based upon oil-in temperature and lubricant flow rate,

bearing temperature can to some reasonable degree of engineering certainty be controlled. Bearing life will be a function

of the resulting bearing temperature and the difference in

temperature between the inner and outer rings. We calculated

bearing lives at three speeds, 30 000, 50 000, and 65 000 rpm

and four hypothetical temperature combinations. These

inner-ring and outer-ring temperature combinations were

154 and 143 oC (310 and 290 oF); 149 and 149 oC (300 and 3

00 oF); 143 and 154 oC (290 and 310 oF) and 138 and 138 oC

(280 and 280 oF). The hotter inner-ring temperature represents

what might be obtained with air-oil mist lubrication while the

hotter outer-ring temperature is more representative of our

results with oil-jet lubrication. The assumption of both rings

being of equal temperature can occur with both methods of

lubrication. The lower ring temperature of 138 oC (280 oF)

represents the lower temperatures obtainable with either lubrication method under the same operating conditions.

The results of the bearing life analysis are summarized in

Table 3. Based upon a lower operating temperature of 138 oC

(280 oF) at 30 000 rpm, the bearing would theoretically have a

56 percent increase in life over the bearing operated at 149 oC

(300 oF). As speed is increased, the life differences decrease. At

65 000 rpm, the difference in life is 9 percent. It can be

reasonably concluded that because gross bearing temperatures

with air-oil mist lubrication are generally higher than that with

oil-jet lubrication, a lower bearing life may occur for a given

operating condition.

Where the bearings were assumed to be nominally at the

same temperature, the oil-jet lubricated bearing at 30 000 rpm

having a lower inner-ring temperature, can have a theoretical

life 47 percent greater than the air-oil mist lubricated bearing.

At 50 000 and 60 000 rpm, the life differences are insignificant.

From this analysis and the measured temperatures, it may be

assumed that where higher bearing temperatures occur with

air-oil mist lubrication, lower bearing life will result.

GENERAL COMMENTS

According to Bloch and Shamim (4), air-oil mist lubrication

systems have been in use since the 1930’s. They describe the

state-of the-art of air-oil mist lubrication systems and discuss

their applications. The economics for using air-oil mist is

compelling based upon acquisition and operating costs as well

as simplicity of design. However, the operating speeds for

which they reported rolling-element bearing operating data

were relatively low, around 200 000 DN. The air-oil mist

lubrication bearing data reported by Schuller and Anderson (5)

was up to 1.2×106 DN. The work reported by Rosenlieb (9,10)

extended rolling-element bearing operating speed with air-oil

mist lubrication to 2.5×106 DN.

The data reported herein for both air-oil mist lubrication and

oil-jet lubrication were for bearing operation where the bearing

temperature and power loss were caused from frictional heat

generation internal to the bearing. It cannot be assumed a priori

that a specific lubrication system will result in lower or higher

bearing temperatures, power loss, and higher or lower lives

than other lubrication systems. The total thermal environment

of the bearing must be considered. As an example, if there is an

external heat source such as a hot section of a turbine engine,

an air-oil mist lubrication system may not have sufficient heat

transfer capacity to maintain the required operating temperature for the bearing. A thermal and heat transfer analysis as well

as a bearing fit-up study needs to be undertaken before a

commitment is made to a specific lubrication system.

Another issue that must be considered is the EHD film

thickness in the ball-race contact. If the calculated lives of

Table 3 were extrapolated to 10 000 rpm without further

calculation of film thickness, it would be concluded that the life

of the bearing would be 3100 hr air-oil mist lubrication and

1500 hr oil-jet lubrication. However, at 10 000 rpm and

temperatures of 138 and 149 °C (280 and 300 °F), the EHD film

thickness is sufficiently low wherein the bearing operates in

the boundary lubrication regime. Accordingly, the calculated

bearing lives for both lubrication systems are the same,

1493 hr. The effects of lubricant additives, which are known to

affect bearing life, were not addressed in this program.

For high-speed turbomachinery applications an issue of

stray (oil) mist must be addressed. As described by Bloch and

Shamim (4), as oil mist passes through the bearing, larger oil

particles are deposited on the bearing surfaces. However,

smaller particles of oil go out with the air. These smaller

particles are called stray mist. Bloch and Shamim (4) report

that 40 to 60 percent of the oil may go out as stray mist.

TABLE 3.—THEORETICAL EFFECT OF

BEARING COMPONENT TEMPERATURE ON

LIFE OF 35-mm BORE ANGULAR-CONTACT

BALL BEARING. THRUST LOAD, 667 N

(150 lbf); NOMINAL CONTACT ANGLE, 24°;

CAGE, SINGLE OUTER-LAND GUIDED

Inner-ring

speed,

rpm

Assumed ring

temperature

°C (°F)

Bearing

life,

hrsa

Inner ring Outer ring

30 000 138 (280)

149 (300)

143 (290)

154 (310)

138 (280)

149 (300)

154 (310)

143 (290)

2049

1315

1576

1072

50 000 138 (280)

149 (300)

143 (290)

154 (310)

138 (280)

149 (300)

154 (310)

143 (290)

899

735

721

703

65 000 138 (280)

149 (300)

143 (290)

154 (310)

138 (280)

149 (300)

154 (310)

143 (290)

274

251

249

246 a

Life factor for effect of elastohydrodynamic film

 thickness included. No other life factors included.
NASA/TM—2001-210462 11

They discuss closed loop air-oil mist lubrication systems that

can capture or recycle stray mist. The ability to capture or

recycle spray mist may be a condition precedent to the application of air-oil mist lubrication systems to high-speed

turbomachinery. An alternate approach for air-breathing turbine engines is to burn the stray oil mist with the fuel.

SUMMARY OF RESULTS

Parametric tests were conducted with an optimized

35-mm-bore-angular-contact ball bearing on a high-speed,

high-temperature bearing tester. The test bearing has a singleouter-ring, land-guided cage with a nominal 24° contact angle.

Both air-oil mist lubrication and oil-jet lubrication systems

were used and compared to speeds of 2.5×106 DN. The oil-in

temperatures were 49 °C (120 °F) and 121 °C (250 °F), for airoil mist and oil-jet lubrication, respectively. The lubricant was

a neopentylpolyol (tetra) ester that met the MIL–L–23699

specification. The following results were obtained:

1. Where the bearing temperature rise is solely caused from

heat internally generated from within the bearing cavity,

the maximum obtainable speed with air-oil mist lubrication is 2.5×106 DN.

2. Lower bearing temperatures and higher power losses are

obtained with oil-jet lubrication than with air-oil mist

lubrication.

3. Bearing power loss is a direct function of oil flow to the

bearing and independent of delivery system. For a given

oil-flow rate, bearing temperature and power loss increase

with increases in speed.

4. Bearing life is an inverse function of temperature, the

difference in temperature between the individual bearing

ring components, and the resultant elastohydrodynamic

(EHD) film thicknesses. Bearing life is independent of the

oil delivery system except as it affects these temperatures.

5. Cage slip increased with increases in speed. Cage slip as

high as 7 percent was measured and was generally higher

with air-oil mist lubrication than with oil-jet lubrication.

REFERENCES

1. Zaretsky, E.V., “Rolling Bearing and Gear Lubrication,”

Tribology for Aerospace Applications, E. V. Zaretsky,

ed., STLE SP–37, Society of Tribologists and Lubrication

Engineers, Park Ridge, IL, pp. 207–323, (1997).

2. Zaretsky, E.V., Signer, H., and Bamberger, E.N., “Operating Characteristics of 120 mm Bore Ball Bearings at

3×106 DN,” NASA TN D–7837, (1974).

3. Signer, H., Bamberger, E.N., and Zaretsky, E.V., “Parametric Study of the Lubrication of Thrust Loaded 120 mm

Bore Ball Bearings to 3 Million DN,” ASME Jour. Lubr.

Tech., 96, 3, pp. 515–525, (1974).

4. Bloch, H.P., and Shamim, A., Oil Mist Lubrication:

Practical Applications, Prentice Hall, (1998).

5. Schuller, F.T., and Anderson, W.J., “Operating Characteristics of 75-Millimeter Bore Ball Bearings at Minimum

Oil Flow Rates Over a Temperature Range to 500 oF,”

STLE Lubr. Engr., 17, 4, pp. 291 –298, (1961).

6. Anderson, W.J., “Extreme-Temperature Bearings, ” Advanced Bearing Technology, E.E. Bisson and W.J.

Anderson, eds., National Aeronautics and Space Administration, Washington, DC, NASA SP–38, pp. 309–370,

(1964).

7. Morrison, F.R., Zielinski, J., and James, R., “Effects of

Synthetic Industrial Fluids on Ball Bearing Performance,”

ASME Jour. Energy Source Tech., 104, 1, pp. 174–181,

(1982).

8. Shamim, A., and Kettleborough, C.F., “Tribological Performance Evaluation of Oil Mist Lubrication,” ASME

Jour. Energy Source Tech., 116, 1, pp. 224–231, (1994).

9. Rosenlieb, J.W., “Microfog Lubrication for Aircraft Engine Bearings,” SKF- AL75TO32 (NASA CR–134977),

SKF Industries, Inc., King of Prussia, PA, (1976).

10. Rosenlieb, J.W., “Emergency and Microfog Lubrication

and Cooling of Bearings for Army Helicopters,” SKF–

AL77TO21 (NASA CR–135195), SKF Industries, Inc.,

King of Prussia, PA, (1978).

11. Jeng, Y.R., and Huang, P.Y., “Temperature Rise of Hybrid Ceramic and Steel Ball Bearings with Oil-Mist

Lubrication,” STLE Lubr. Engr., 56, 12, pp. 18–23, (2000).

12. Pinel, S.I., Signer, H.R., and Zaretsky, E.V., “Design and

Operating Characteristics of High-Speed, Small-Bore,

Angular-Contact Ball Bearings,” STLE Trib. Trans., 41,

4, pp. 423–434, (1998).

13. Pinel, S.I., and Signer, H.R., “Development of a HighSpeed, Small Bore Bearing Test Machine,” NASA

CR–135083, (1976).
This publication is available from the NASA Center for AeroSpace Information, 301–621–0390.

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August 2001

NASA TM—2001-210462

E–12458

WU–253–02–98–00

17 Rolling-element bearings; Ball bearings; Life prediction;

Lubrication; Oil-jet lubrication; Air-oil mist lubrication

Unclassified -Unlimited

Subject Category: 37 Distribution: Nonstandard

Prepared for the 56th Annual Meeting sponsored by The Society of Tribologists and Lubrication Engineers, Orlando,

Florida, May 20–24, 2001. Stanley I. Pinel, Pinel Engineering, Fullerton, California; Hans R. Signer, Signer Technical

Services, Fullerton, California; Erwin V. Zaretsky, NASA Glenn Research Center. Responsible person, Erwin V. Zaretsky,

organization code 5900, 216–433–3241

Stanley I. Pinel, Hans R. Signer, and Erwin V. Zaretsky

Comparison Between Oil-Mist and Oil-Jet Lubrication of

High-Speed, Small-Bore, Angular-Contact Ball Bearings

Parametric tests were conducted with an optimized 35-mm-bore-angular-contact ball bearing on a high-speed, hightemperature bearing tester. Results from both air-oil mist lubrication and oil-jet lubrication systems used to lubricate the

bearing were compared to speeds of 2.5×106 DN. The maximum obtainable speed with air-oil mist lubrication is

2.5×106 DN. Lower bearing temperatures and higher power losses are obtained with oil-jet lubrication than with air-oil

mist lubrication. Bearing power loss is a direct function of oil flow to the bearing and independent of oil delivery

system. For a given oil-flow rate, bearing temperature and power loss increase with increases in speed. Bearing life is an

inverse function of temperature, the difference in temperature between the individual bearing ring components, and the

resultant elastohydrodynamic (EHD) film thicknesses. Bearing life is independent of the oil delivery system except as it

affects temperature. Cage slip increased with increases in speed. Cage slip as high as 7 percent was measured and was

generally higher with air-oil mist lubrication than with oil-jet lubrication.