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Research Papers: Gas Turbines: Heat Transfer

Active Outer Ring Cooling of High-Loaded and High-Speed Ball Bearings

[+] Author and Article Information
Francois Cottier

MTU Aero Engines,
Munich 80995 Germany

Peter Gloeckner

FAG Aerospace,
Schweinfurt 97421 Germany

Klaus Dullenkopf

University of Karlsruhe,
Karlsruhe 76128 Germany

Contributed by the Heat Transfer Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received February 26, 2013; final manuscript received March 15, 2013; published online June 24, 2013. Editor: David Wisler.

J. Eng. Gas Turbines Power 135(8), 081902 (Jun 24, 2013) (8 pages) Paper No: GTP-13-1062; doi: 10.1115/1.4024257 History: Received February 26, 2013; Revised March 15, 2013

Bearings for aero engine applications are subjected to a high thermal impact because of the elevated speeds and loads. The high rate of heat generation in the bearing cannot be sustained by the materials used and, in the absence of lubrication, will fail within seconds. For this reason, aero engine bearings have to be lubricated and cooled by a continuous oil stream. The heat that is generated in the bearings through friction is transferred into the oil. Oil itself has limited capabilities and can only remove heat as long as its temperature does not reach critical limits. When the critical limits have been reached or even exceeded, the oil will suffer chemical decomposition (coking) with loss of its properties and subsequently cause a detrimental impact on the rotating machinery. Oil is normally transferred into the bearings through holes in the inner ring, thus taking advantage of the centrifugal forces due to the rotation. On its way through the bearing, the oil continuously removes heat from the inner ring, the rolling elements, and the bearing cage until it reaches the outer ring. Since the oil has already been heated up, its capability to remove heat from the outer ring is considerably reduced. The idea to provide the bearing with an “unlimited” quantity of oil to ensure proper cooling cannot be considered, since an increase in the oil quantity leads to higher parasitic losses (churning) in the bearing chamber and increased requirements on the engine's lubrication system in terms of storage, scavenging, cooling, weight, etc., not mentioning the power needed to accomplish all these. In this sense, the authors have developed a method that would enable active cooling of the outer ring. Similar to fins, which are used for cooling electronic devices, a spiral groove engraved in the outer ring material would function as a fin body with oil instead of air as the cooling medium. The number of “threads” and the size of the groove design characteristics were optimized in a way that enhanced heat transfer is achieved without excessive pressure losses. An experimental setup was created for this reason, and a 167.5-mm pitch circle diameter (PCD) ball bearing was investigated. The bearing was tested with and without the outer ring cooling. A reduction of 50% of the lubricant flow through the inner ring associated with a 30% decrease in heat generation was achieved.

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References

Figures

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Fig. 1

Ball bearing with the outer ring groove

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Fig. 2

Details of the bearing and of the spiral groove

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Fig. 3

The instrumentation setup

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Fig. 9

Parameter D0 introduced in Eq. (2)

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Fig. 10

Comparison between the Flouros/Hirschmann and the Joshi and Shah friction correlation for flows in spiral ducts

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Fig. 8

The introduction of outer ring cooling can considerably reduce the oil flow rate and the heat rejection from a bearing

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Fig. 7

The evolution of the outer and inner ring temperatures (TOR, TIR) and of the heat rejection into the oil as a function of the oil flow at 17,000 rpm and 80-kN axial load

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Fig. 6

The evolution of the scavenge temperatures on both sides of the bearing as a function of the oil flow at 17,000 rpm and 80-kN axial load

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Fig. 5

(a) The pumping effect as a function of the axial load at 17,000 rpm. (b) The pumping effect as a function of the rotor speed at 80-kN axial load.

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Fig. 4

The pumping effect as a function of the oil flow through the inner ring at 17,000 rpm. The mean average scavenge flow through the LHS is about 75%.

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Fig. 11

Comparison among different Nusselt correlations at Prm = 65 and L = 3070 mm

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Fig. 17

Comparison between the temperature distributions in the hub/outer ring plane with and without the outer ring cooling

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Fig. 18

Almost 28 K of temperature gradient is achieved in a fin with only 2.3 mm of depth

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Fig. 12

The crop factor as a function of NTU

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Fig. 13

Flow chart recommending the design procedure

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Fig. 14

The models for the inner and outer rings, including the outer ring hub. The dark zones are the contact zones for the rolling elements.

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Fig. 15

The temperature distribution in the inner ring. The areas around the slots for the oil have the lowest temperature, whereas the contact zones with the rolling elements have the highest temperatures (darkest areas). The average computed temperature is 172 °C, whereas the measured average is 176 °C.

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Fig. 16

The model of the outer ring with hub and helical duct

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