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Thrust Cone Bearings Provide Increased Efficiency for Helical Gear Units at Moderate Speed Levels



Thrust cone bearings are an elegant option to handle the axial forces generated by the torque transmission in helical-toothed gear stages. They have proven as an efficient and reliable bearing concept for integrally geared compressors but are nearly unknown in other fields of gearbox engineering. The presented investigations consider three aspects which appear relevant to extend the field of possible applications for thrust cones towards gearboxes constructed with roller bearings. Based on simulations and experiments design parameters were identified, which enable a significant reduction of the necessary velocity for full film lubrication. For a single stage test gearbox noticeable increases in efficiency were achieved by replacing tapered roller bearings with a combination of thrust cone and ball bearings, especially during partially loaded operation. The resistance to wear and the determination of limits for the bearable loads under mixed friction conditions for various thrust cone design configurations are investigated in a third test series. It appears that the few limit values known so far might be exceeded significantly for future applications.


Helical gears are a common solution to reduce noise and increase the transmittable torque in the construction of gearboxes. Unfortunately the pair of contact forces between the meshing tooth flanks is not perpendicular to the axis of rotation of the gear shafts, due to the helix angle. Transmitting torque between pinion and gear leads to an axial force component, which usually has to be transferred through the gears, the shafts and axial bearings into the housing of the gearbox (Fig. 1, left) or offset by the use of double-helical gears.

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Figure 1 Axial force components in helical gear boxes: with conventional bearing concept (left) and thrust cone bearing setup (right) (Ref. 6).

Thrust cone bearing concept. Figure 1 (right) presents an alternative-bearing concept, i.e. — the “thrust cone” bearing. Conical rims — denoted as thrust cones — are attached to both sides of the pinion and flank the opposing wheel. Their conical shape and the contact surfaces ground to the sides of the wheel lead to a narrowing gap in the overlapping area. Lubrication fluid, sticking to the surfaces, is transferred into this gap and generates a hydrodynamic pressure film that separates thrust cones and the contact surfaces on the wheel. The axial force generated on the pinion tooth is transferred through the thrust cone and the fluid film onto the conical contact surface of the wheel; here it meets the axial force component created on the wheel’s tooth. Since both force components obtain the same value — but with opposing directions — they “cancel” each other and no axial force is transferred to the shafts or the housing (Fig. 1, right). This load reduction enables a lighter construction for the machine components; the pinion shaft can be designed without an axial bearing, while the axial bearing on the wheel shaft operates only as positioning, i.e. — without load from the helical gear pair.


The frictional losses in sliding bearings rise proportionally to the square of the differential velocity between the contacting surfaces. In a typical axial sliding bearing, the differential velocity equals the circumferential velocity of the running surface. In a thrust cone bearing, a lower differential velocity occurs between the contacting surfaces since their contact region is located close to the pitch point of the gear pair (Fig. 2).

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Figure 2 Contact velocities for two rotating cones: close to the pitch point favorable cinematic conditions for hydrodynamic lubrication occur (low differential velocity and high sum of surface velocities) (Ref. 6).

Aspects concerning an expanded field of application. Currently the main application for thrust cone bearings is in the field of integrally geared compressors (IGCs), which are characterized by a very high rotational speed on the pinion shaft (more than 10,000 rpm) and a nearly constant torque load at their point of operation. Since the first patent in 1924 (Ref. 2) the thrust cone bearing has proven to be an appropriate, alternative bearing concept for helical gear pairs. (Langer, Ref. 4) stated that a reduction of bearing-related frictional losses for thrust cone concepts to 10–20% compared to classical concepts based on tilting pads. Nevertheless, in special operational situations, such as emergency shutdown sequences, difficulties occur if the hydrodynamic carrying capacity of the lubricant film is not sufficient to separate the contact surfaces. Apart from IGCs the application of thrust cones in modern engineering is limited and nearly negligible.

To improve the reliability of thrust cone bearings and widen the field of possible applications, the following three main topics are within the scope of a research project currently conducted at the Institute for Mechanical Engineering (IMW):

Low-speed, full-film lubrication. Aiming for higher gearbox efficiency, it is a crucial condition that the friction losses in the thrust cone bearing are lower than the reduction of losses in the bearings of the housing. Since the friction coefficient under full-film lubrication is significantly lower than in the mixed-friction regime, the importance of the required reduction of the transition velocity becomes obvious.

Proof of efficiency. In ICGs, thrust cone bearings are usually combined with journal bearings where the axial load compensation supersedes one axial bearing. In gearboxes based on roller bearings, the benefit in efficiency is gained by a change of bearing types. Instead of tapered roller bearings, which are distinguished by their high axial load-carrying capacity but generate a relatively high-energy consumption, more efficient ball bearings might be used if axial loads are compensated by thrust cones. To prove that the suggested change in bearing types outweighs the additional friction in the thrust cone contact, a comparative examination of gearbox efficiency for both concepts is presented.

Determination of bearable load under mixed-friction conditions. Even though the transition velocity is reduced, there will remain situations (starting or breaking maneuvers, for example) with insufficient speed for full-film lubrication. A successful thrust cone design must safely withstand these mixed-friction situations during the product’s life cycle. Unfortunately, at this writing there is as yet no available information on bearable loads for thrust cone bearings in open literature. To enable a wider use of thrust cone bearings as a resource-efficient machine element in gearboxes, a description of possible design influences on limiting load values under mixed-friction conditions is required.

Reduction of the Required Velocity for Full-Film Lubrication

Since full-film lubrication achieves efficient operation and nearly eliminates wear effects on the contacting surfaces, research activities were initially focused on influences on the transition velocity. In general the fluid film in a thrust cone bearing increases with the rotational speed and reduces with additional load — but for a certain operation point (combination of load and speed) various thrust cone designs generate different fluid film thicknesses. To predict the effect of design parameters on film thickness, a hydrodynamic fluid film simulation was developed. The algorithm — inspired by the work of (Barragan de Ling, Ref. 1) — allows solving the Reynolds differential equation for a thrust cone bearing, calculates the hydrodynamic pressure distribution, and determines the minimum gap size between the elastically deformed contact surfaces.

Figure 3 illustrates some design variations for thrust cone bearings, influencing the transition behavior. Besides variations of cone angle and slip value (depending on the distance between pitch point and contact surface), macroscopic shape variations for the running surface geometry were within the scope of our examinations.

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Figure 3 Design variations for thrust cone bearings (Ref. 6).

To validate the predictions made by simulation, experiments on the thrust cone test bench (Ref. 7) were carried out. For various thrust cone specimens, representing the shape designs used in the simulations earlier, the transition velocities at different load steps were determined. In the full-film lubrication region, contact surfaces are completely separated by the lubrication fluid. Due to its low conductivity, high electric resistance can be observed between the specimens. If the velocity is reduced the film thickness shrinks for a constant axial load. A drastic drop of the electrical resistance is notable when first metallic contact occurs between the peaks of the rough surfaces. The current combination of load and velocity is regarded as a transition point to the mixed-friction regime. Some of the determined transition points are plotted (Fig. 4) as an example of the experimental work. The data illustrates borderlines between mixed- friction and full lubrication for flat thrust cones with different inclination angles and slip values. Since low- transition velocities are desired for early full-film lubrications, optimal design configurations can be found in the lower region (Fig. 4). In accordance with simulations, low slip values and cone angles lead to an earlier separation of contact surfaces; nevertheless, a cone angle greater than zero is necessary to generate even the slightest hydrodynamic lubrication.

Figure 4 shows that a small geometric variation in the bearing design can have a great influence on transition behavior. The transition velocity for the best design (inclination angle: 0.5°; slippage: 10%) is about 50% lower than for the worst in the Fig. (inclination angle: 1.0°; slippage: 20%). Assuming constant acceleration during a starting maneuver, the distance to be run under mixed-friction conditions is reduced by the factor 4.

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Figure 4 “Border lines” for the transition between mixed friction and full film lubrication for different flat thrust cones based on the experimental data.

What’s more, it becomes obvious that the transitions observed in the test occur within a speed range that is significantly lower than the operational velocities for thrust cone bearings in ICGs (> 100 m/s), and even relevant for gearboxes running with roller bearings.

Gearbox Efficiency in Dependency of the Bearing Concept

To investigate the thrust cones’ ability to reduce friction losses in gearboxes based on roller bearings, a test gearbox (Fig. 5) has been set up. The gearbox is driven by an inverter-fed, asynchronous machine; a torque load can be applied by an adjustable mechanical break coupled to the output shaft. Input and output torque and the rotational speed are recorded via measurement shafts with integrated rotary encoders. Both power input and output can be calculated from the captured data. The difference between input and output is regarded as system losses of the gearbox and the reciprocal quotient as its efficiency. The system losses are the sum of losses caused by several gearbox elements. To evaluate the influence of the bearing concept on the system losses, the gearbox can be equipped with either a set of tapered roller bearings in O-arrangement, or with a combination of thrust cone and ball bearings. All other components contributing to the system losses — gearing, seals, oil level, etc. — are kept constant during the tests.


2024-08-22