When the load-carrying capacity of a bearing is insufficient, it can lead to premature wear, increased friction, and even bearing failure, which may cause significant operational issues and downtime in various machinery and mechanical systems. Fortunately, there are several methods and solutions available to increase the load-carrying capacity of bearings, ensuring better performance, longevity, and reliability. Below are some common and effective strategies:
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One of the most direct and simple approaches to increasing the load-carrying capacity of a bearing is to enlarge its size. By increasing the bearing's diameter, the surface area available for load distribution is expanded, which helps to share the load more effectively. A larger bearing can handle greater forces without experiencing undue stress on its components.
Additionally, adjusting the ratio of the bearing's width to its diameter can also improve its load capacity. Increasing the bearing width can allow for a thicker film of lubricating oil, which reduces the chances of metal-to-metal contact, leading to lower friction and wear. This approach can be particularly useful in applications where space allows for larger bearings without compromising the overall design.
Turbine Bearings
Lubrication plays a critical role in reducing friction and wear in bearings. One way to enhance the load-carrying capacity is by increasing the viscosity of the lubricating oil. Higher-viscosity oils form thicker lubrication films between the bearing's moving parts, which can help to prevent metal-to-metal contact under heavy loads. This thicker oil film ensures that the load is distributed more evenly and reduces the chances of surface degradation or damage.
However, it is essential to balance the viscosity of the oil with the operating conditions, as overly viscous oils can increase friction at low speeds and impair the bearing’s efficiency. Therefore, selecting the appropriate viscosity for specific conditions, such as load, speed, and temperature, is crucial.
In some cases, the solution may lie in upgrading to a bearing that is specifically designed to handle higher loads. Bearings such as circular bearing bushes and tilting pad thrust bearings are excellent choices for heavy-duty applications. Circular bearing bushes are typically used in cases where the bearing needs to accommodate large radial loads, while tilting pad thrust bearings are ideal for applications that require the handling of axial loads.
Tilting pad thrust bearings, for example, consist of multiple pads that can tilt in response to the load, which allows for an even distribution of pressure across the bearing's surface. This design helps to significantly increase the load-carrying capacity, especially in high-speed and high-load applications such as turbines and large motors.
tilting pad journal bearing
Hydrostatic and hydrodynamic composite bearings are advanced bearing solutions that use high-pressure oil to create a stable and reliable lubrication film between the bearing surfaces. In hydrostatic bearings, oil is supplied at a high pressure through an external source, which helps to separate the bearing surfaces completely, preventing direct contact and reducing wear and friction. This solution is particularly effective for applications requiring heavy loads and low speeds.
Hydrodynamic composite bearings combine the benefits of both hydrodynamic and hydrostatic principles. These bearings use the motion of the shaft to generate the necessary oil film for lubrication, while also incorporating high-pressure oil at certain points to maintain a stable lubrication layer. This combination can significantly improve the bearing's load-carrying capacity, especially in applications where the load varies or is extremely high.
Bearings made from PTFE (polytetrafluoroethylene) composite materials offer excellent performance in terms of load capacity and wear resistance. PTFE is a high-performance plastic known for its low friction, high wear resistance, and ability to withstand harsh operating conditions. Bearings made from PTFE composites can carry heavier loads without experiencing excessive wear, making them an ideal choice for applications that require long-lasting, high-performance bearings.
PTFE composite materials are often used in self-lubricating bearings, as they reduce the need for external lubrication. This makes them ideal for environments where maintaining lubrication is challenging, such as in high-temperature or underwater applications.
To address insufficient load-carrying capacity in bearings, there are various methods available, each tailored to different operational needs and environments. Whether by increasing the bearing size, enhancing lubrication, upgrading to higher-capacity bearings, or adopting advanced bearing technologies such as hydrostatic or hydrodynamic composite bearings, each solution plays a vital role in ensuring that the bearing can withstand heavy loads without compromising performance or durability. Choosing the right solution depends on the specific requirements of the application, including load, speed, temperature, and space constraints, ensuring optimal performance and longevity of the bearing system.
Oil-whirl instability in rotor bearings was discovered by General Electric engineer Burt Newkirk in the s. Since then, much has been written about design methods to reduce its harmful effects. Yet it still remains troublesome and of great concern, particularly in lightly loaded journal bearings in turbines, compressors, and pumps; in large vertical motors and generators; and in many similar machines.
Here’s a brief review of the physical nature of oil whirl, some ways to reduce or eliminate its effects, and alternative designs to avoid it altogether.
Three types of shaft vibration are particularly pertinent to fluid-film journal bearings:
Unbalance vibration is almost always caused by a lack of balance in the rotating mass supported by the bearings. This type of vibration can be alleviated or eliminated by carefully balancing the rotor.
Half-frequency oil whirl, a type of rotary-shaft motion, is caused by a wedge of oil film traveling around the bearing circumference at an average velocity of half the shaft’s surface speed. Amplitude of this rotor whirl often approaches 50 to 100% of the total clearance and threatens machine performance.
Oil whip, a potentially catastrophic vibration, occurs when whirl frequency coincides with the natural frequency of the shaft at rotational speeds two or more times the rotor’s natural frequency. This might be thought of as a stabilized version of half-frequency whirl. Its constant frequency is half the first natural frequency of the shaft for speeds ranging up to about three times the natural rotor frequency.
The effects of these three conditions are shown graphically in Rotor-vibration frequency.
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Complexities associated with the oil film, bearing geometry, temperature, viscosity, supporting structure, and the like make it difficult to predict rotor instability. In fact, different machines may respond differently with the same bearings, and secondary effects may influence the type of vibration.
For example, as speed increases unbalance may delay a shift from simple rotational frequency to expected half-frequency oil whirl. On the other hand, at speeds up to around three times the natural rotor frequency, unbalance vibration takes over and amplitude rises to the maximum possible within the bearing clearance.
What causes whirl? The answer involves the interaction of external forces on the bearing and internal oil-film pressure generated within the bearing clearance.
Ideally, entrained oil in the bearing clearance circulates with an average velocity of about half the shaft surface speed. But in practice, it is between 40 and 48% of shaft speed. With a load on the bearing, the shaft center becomes eccentric and the shaft surface rides on the crest of an oil film.
Now, with normal radial displacement of the shaft, oil pressure on the upside of the close clearance zone will be higher than on the downside. The resulting tangential force in the direction of rotation tends to induce “forward whirl” of the journal. With this whirl, the increasing tangential force reduces the clearance further to produce an increasingly higher destabilizing tangential force. Field experience has shown this whirl frequency is typically less than 2% below half the rotor speed.
Let’s look at some highly simplified features of oil whirl. The graphic, Journal bearing stiffness and damping, represents an idealized rotor of weight 2W supported on two identical bearings. The load on each bearing is W, and the bearing center is displaced to a steady operating position as shown. One can draw a line connecting the shaft and bearing centers, and the load line, to form an attitude angle ø. It’s unique to the bearing and operating condition.
In general, the lowest natural frequency is No = ωo/2π, where ωo = (k/m)0.5 with k representing shaft stiffness and m the rotor mass. While the equivalent spring constant for vertical displacement is not linear over any extended range, a value KXX can be taken as the differential displacement in the vertical direction (X) with a differential change in W (as with a rotor imbalance or varying external force). With no rotor flexibility and negligible damping, the bearing vertical natural frequency becomes, as for any simple spring-mass system:
Similarly, one can define damping coefficients CXX and CYY in the vertical and horizontal directions. For other than pivoted-pad bearings, cross-coupling coefficients must also be introduced (KXY, for instance) as a spring constant for motion in the Y direction, for varying vertical load W in the X direction. Forces induced by the cross-coupling coefficients tend to drive the rotor into an orbit. Typically for light loads in fixed-arc bearings with cross-coupling, this orbit frequency is close to half the journal rotational speed.
If damping and load are not sufficient to suppress this cross-coupling force, the rotor may break violently into an uncontrolled whirl orbit, or rotor whip, with an amplitude closely matching the full internal clearance in the bearing.
Stability characteristics of a rotor carried on a full 360° journal bearing can be estimated from the Stability limits graph. These results are based on the so-called short bearing approximation valid for bearings with L/D ≤ 1 and apply only for laminar flows.
Greater shaft flexibility (lower Ks) tends to decrease stability. Kscan be approximated from simple beam theory by dividing the rotor weight by the midspan deflection resulting from that weight. (A more-detailed analytical background and oil-film response coefficients for fixed-arc and pivoted-pad bearings are provided by “Journal and Thrust Bearings” by A. Raimondi and A.Z. Szeri; and “Fluid Film Lubrication” by H.J. Sneck and J.H. Vohr. Both can be found in the CRC Handbook of Lubrication, Vol. II, CRC Press, Boca Raton, Fla., .) More-recent computer routines using numerical solutions deal with the effect of temperature on the bearing’s dynamic behavior.
Here’s an example of how to predict behavior. Determine whirl stability for a 5,000-lb horizontal rotor in its cylindrical bearings with the following characteristics: diameter D = 2R = 6 in.; length L = 3 in.; radial clearance C = 0.006 in.;
N = 60 rev/sec (3,600 rpm); lubricant viscosity µ = 2 ×
10-6 lb-sec/in.2 (reyns); and rotor stiffness Ks= 10 × 106 lb/in.
Load or rotor mass m = W/g = 5,000/386 =13.0 lb-sec2/in.
Unit load P = W/(DL)= (5,000/2)/(6 × 3) = 139 psi.
Sommerfeld number S = (R/C)2 (μN/P) = (3/0.006)2 (2 × 10-6 × 60/139) = 0.216.
Bearing characteristic number S(L/D)2 = 0.216(3/6)2 = 0.054.
(C/W)Ks = (0.006/2,500) × (10 × 106) = 24; and
(C/W)mω2 = (0.006/2,500) × 13.0 × (2π × 60)2 = 4.43.
Referring to the Stability limits graph, because the coordinate point S(L/D)2 = 0.054, (C/W)mω2 = 4.43 lies below the curve for (C/W)Ks = 10 and the rotor will be free of oil-whirl instability.
If the speed were raised to 7,200 rpm, however, repeating the above calculations shows that S(L/D)2 = 0.108 and
(C/W)mω2 = 17.72, which falls in the unstable region.
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