Bearing lubrication device

Bearing lubrication device.

Grease rings are widely used as conduits for conveying oil or other lubricant from a reservoir to moving parts such as bearings, shafts and the like. In operation, the grease ring is normally placed freely around the shaft and, when the latter rotates, it also rotates in contact with this shaft.

Lubricant is conveyed from a housing or reservoir to the shaft through the contours or grooves of the grease ring, as well as by frictional attraction as this ring moves through the reservoir. Lubricant is deposited on the shaft or other component by the forces of gravity, friction and centrifugal inherent in operation. Under conditions of slow rotation, the forces of gravity and friction generally distribute a sufficient amount of lubricant; however, at higher speeds which can be as high as 915 to 1,220 m / minute, the grease ring moves too quickly for gravity to disperse the oil or the centrifugal force exerted on the ring 25 and the oil is too much to overcome and the oil remains on the ring or is thrown out of the field of rotation. Therefore, the lubricant. does not achieve the desired surface, resulting in premature wear and possibly failure of the shaft, bearing, grease ring or other associated members.

The rotation of the ring depends on a propulsive force exerted between the rotating shaft and this ring. As the speeds increase, a fluid film forms and the driving force is formed.

<img file="LU85569A1_D0001.tif" />

This film of lubricant is transmitted to the ring. In many respects the situation is analogous to that in a floating ring bearing and, if a direct drive mechanism is not provided, slippage occurs.

Previous attempts to create a higher coefficient of friction and hence a more positive drive mechanism have focused on changing the geometry of the inner circumference of the ring. The factors opposing the rotation of the ring are the excessive friction exerted on the lower part of the ring, which is immersed in the lubricant reservoir, the force required to lift the lubricant out of the reservoir and bring it. at the top of the journal and the frictional resistance exerted on the ring by fixed, press-fit surfaces such as the sides of the annular groove in the bearing. Among other factors affecting the distribution of the lubricant are the composition of the ring and the viscosity of the lubricant used in the bearing. Additionally, since a conventional grease ring sits on the top surface of the shaft during operation and during periods of non-use, significant wear results from contact only. At rest, most of the lubricant returns to the reservoir and there is very little lubricant protection for the starting operation. Therefore, before the lubricant film is restored, premature wear of the shaft, ring, bearings and other associated components can occur, which in turn leads to costs. repair and replacement, not to mention that it also results in loss of operating time.

Accordingly, one of the main objects of the present invention is to improve the lubricity of the lubricating rings, thereby increasing the possibilities and the capacity of the thrust bearings and the bearing shells by providing a bearing lubricating device. comprising an assembly of a grease ring and a cantilevered oil blade, An assembly in which this cantilever blade scrapes lubricant from the contours or grooves of the grease ring by directing and depositing the lubricant in the desired areas around the shaft, bearings and ring itself.

Another object of the present invention is to reduce the rotational speed of the grease ring, thereby ensuring greater distribution of the lubricant by utilizing the effective braking property provided by the action of the cantilevered oil blade. against the flow of lubricant conveyed by this ring from the reservoir, while increasing the lubricant distribution due to the ring configuration and the divergent wedge configuration of the oil blade.

Another object of the present invention is to provide a spring support for the grease ring and to prevent the excessive rocking movement which characterizes the types of grease rings of the prior art, by providing a means to minimize contact. between the ring and the shaft before and during operation, thus minimizing the wear, at start-up, of the grease ring, the shaft and the bearings, while prolonging the useful life of these components.

j?

Still another object of the present invention is to provide a cantilevered grease ring and oil blade assembly which can be used with most or all devices in which grease rings are commonly employed. conventional, this assembly being of an economical realization and use.

These various objects, as well as others, are achieved by the present invention which relates to a bearing lubricating device for use in bearing bushes lubricated by a ring and the like, comprising a rotating shaft, a bearing surface. and a lubricant reservoir, this device comprising a generally circular annular member for conveying the lubricant from the reservoir to deposit it on the shaft and the bearing surface. The annular member rotates with the shaft and means is provided to separate the lubricant from the ring, thus providing a greater distribution of lubricant than that obtainable with conventional grease rings. For bidirectional bearing shells, a second lubricant collecting means may be added opposite the first means in order to direct the lubricant onto the shaft and the bearing surfaces, as well as into the feed grooves of the bearing pins, when these grooves are present.

Various other objects and advantages of the present invention will become apparent on reading the description given below with reference to the accompanying drawings in which:

Figure 1 is a side elevational view partially in cross section of a support bearing type bearing bush assembly including the following lubricating device

JF the present invention disposed around the shaft of the bearing assembly;

FIG. 2 is a partial perspective view of the device for lubricating bearings according to the present invention, this device being, in this case, installed in a bearing of the support bearing type, part of the structure of the bearing being pruned in order to show the orientation of the grease ring and the cantilevered oil blade with respect to the shaft;

Figure 3 is an enlarged perspective and partially cross-sectional view of the grease ring and the cantilevered oil blade according to the present invention, this figure showing the contours of the ring, while the section is taken in circle 3 of figure 2;

Figure 4 is a partial, schematic and graphical representation of the various positions taken by the cantilevered oil blade for a shaft rotatable in one direction only;

Figure 5 is a partial, schematic and graphical representation of another embodiment of the present invention, this figure showing the various positions taken by the cantilevered oil blade on one side of the bearing structure and a blade separate oil collector facing this cantilever blade for a shaft that can turn in two directions;

Fig. 6 is a graph illustrating the relationship between shaft speed and oil distribution for a single grease ring, as well as for a grease ring with a cantilevered oil blade;

Figure 7 is a graph illustrating the relationship between shaft speed and distribution.

<img file="LU85569A1_D0002.tif" />

tion of oil for three lubricating rings each comprising a cantilevered oil blade, these rings comprising grooves having different depths;

Fig. 8 is a graph illustrating the relationship between the speed of the shaft, the speed of the ring and the distribution of oil for the grease ring alone constituting a part of the present invention;

Fig. 9 is a graph illustrating the relationship between shaft speed and oil distribution for three lubricants of different viscosity used in accordance with the present invention; and Fig. 10 is a graph illustrating the relationship between the speed of the shaft and the oil distribution for a shaft rotatable in two directions and for a bearing shell using the grease ring, the oil blade in. overhang and oil collecting blade shown in figure 5.

Referring now more specifically to the accompanying drawings and, in particular, to Figure 1, reference numeral 10 generally designates the bearing lubricator according to the present invention. In this case, as illustrated, the device is installed in a bearing shell 12, although its application is by no means limited to this arrangement. The assembly can normally be used whenever conventional grease rings are employed for lubrication purposes and in a variety of different devices. In normal operations with bearings of the type illustrated, the grease ring 19 is freely placed around a rotary shaft 14 with which it rotates as will be explained below. The grease ring rotates in

<img file="LU85569A1_D0003.tif" />

an annular groove 16 through a lubricant reservoir 18 and, during rotation, it drives the lubricant upwards to deposit it on the shaft and the bearing surfaces.

Figure 1 is a partial cross section of the grease ring 19 forming part of the present invention. One of the limiting factors in achieving greater oil distribution and stable operating conditions with grease rings is the configuration of the outer surface. According to the present invention, it has been found that the relative angle of the angular faces 20 together with the length of the vertical faces 22 exerts the greatest influence on the distribution of oil as illustrated in Fig. 3.

When the angle of the faces 20 approaches 0 °, the lateral friction of the ring in the annular groove 16 approaches the maximum possible. As a result, the ring operates irregularly due to the increased lateral friction and the oil distribution is reduced due to insufficient speed of the ring. As the angle of the faces 20 increases thereby shortening the length of the faces 22, the oil distribution increases accordingly and, under the effect of rotational forces, the lubricant is sprayed away from the ring. as a splash or spray. By experience, it has been found that the optimum angle for the angular faces 20 is about 30 ° regardless of the diameter of the ring or the depth of the inner groove referred to herein as reference numeral 24.

When the journal is operating at low speeds, the grease ring follows it and they both have roughly the same peripheral speed.

As the speed of shaft 14 increases, a transition point is reached at which a hydrodynamic film of lubricant begins to form, while significant slip occurs and a large reduction in speed is observed. of 5 the grease ring. At this transition point, it is considered that the speed of the ring is its primary speed vis-à-vis the speed of the journal, which is denoted by in Figure 8. The relation at this point is dUr / dUs = 0 for Ur = where is the primary speed of the grease ring, Ur is the superficial speed of the inner diameter of the ring 19 and Us is the superficial speed of the journal.

The primary speed of the grease ring 15 is a combined function of its weight, shape, projected contact areas, journal speed, lubricant viscosity and localized temperature. As the speed of the journal increases, thereby increasing the speed of the ring beyond the primary speed, a hydrodynamic film of lubricant is precisely established between the ring and the journal. The point at which the actual rotational speed of the ring strikes a balance between the propulsive force in the contact area between the ring and the journal, and the resistance force of the friction of the lubricant on the ring, is called the speed. secondary or N₂· This point is also illustrated in figure 8 and the relation is expressed by dUr / dUs = 0 where Ur = Ng.

Secondary speed is also a function of many parameters, including journal speeds, oil viscosity, ring immersion level and ring shape. For example, the greater the length of the vertical face 22, the greater the secondary speed N₂ is weak.

Beyond the secondary speed, the distribution of the lubricant increases very rapidly with an increase in the speed of the ring. Likewise, as the speed of the journal continues to increase, the ring is driven entirely by hydrodynamic action through the intervention of a thicker film of lubricant. As more lubricant is withdrawn from the reservoir, the friction decreases as a result of the reduced dynamic immersion level of the ring in the lubricant reservoir as a result of the faster rotation. During tests which have been carried out, at a particular speed of a shaft, various rings have undergone excessive vibrations beyond the secondary speed 15. The vibration modes could be easily observed and they were translational, conical and oscillating modes, the vibrations being initiated with an oscillating mode. The amplitude of the vibrations increases as the speed of the shaft increases.

This particular ring speed is considered to be its tertiary speed denoted by in Figure 8. The tertiary ring speed is believed to be the first critical rigid body speed of that ring.

As the journal speed increases beyond the tertiary speed and arrives in the unstable zone, the unstable movement of the ring triggers the projection of the lubricant both out of that ring and out of the journal. This spraying and spraying becomes so vigorous that the lubricant delivery decreases rapidly as shown in Figure 8. Beyond the tertiary speed N „, whatever the speed O of the journal, the speed of rotation of the ring remains constant or decreases. Various specific factors influence this tertiary speed, in particular the shape of the ring, the configuration of the bore of the latter which strongly controls the hydrodynamic rigidity, the weight or the mass of the ring, as well as its diameter; for example, a large ring has a lower tertiary speed. The effects of changes in lubricant viscosity on ring speed and lubricant distribution were investigated using SAE 10, 20, and weight lubricants, and viscosity was found to influence primary speeds. and secondary of the ring; however, the tertiary rate was found to be independent of viscosity.

Different materials can be used for the manufacture of the ring 19, including brass, Muntz (60% Cu, 40% Zn) and bronze (SEA-660). Tests carried out on these materials using SAE 10 lubricant at 37.8 ° C and submerging the ring to a level of 15% of its diameter, indicated that the bronze was able to achieve a distribution of. oil approximately 10% higher than that of other materials tested. Wear property testing consisting of 30,000 start / stop cycles and 7,200 hours of continuous operation at 1,800 rpm with SAE 10 lubricant indicated less wear with the brass ring, however, differences were slight.

Referring again to Figure 2, as shown, the grease ring 19 is disposed around the shaft 14. The shaft may rotate in a bearing or a seal 40 which may be no. any suitable type and, in the illustrated embodiment, the rotation is in the direction of the arrow indicated. A means for separating the lubricant from the ring or the cantilevered oil blade 42 is attached to the packing by means of suitable elements such as screws 46.

The blade 42 has a diverging wedge-shaped configuration and is mounted in a one-way bearing such that the direction of rotation of the shaft 14 is directed towards and into the free end 48 of this blade. The free end 48 is disposed in the groove 24 of the ring 19 and the blade may be made of any suitable material, for example, steel sheet. The design was experimentally optimized and the sheet had a thickness of about 0.5 mm; an arc of about 70 ° has been found to provide optimum performance for any combination of ring and journal. The curved sheet is subjected to a prestress of 10% of the weight of the ring and assumes approximately the position indicated by 50A in Figure 4 when the apparatus is at rest, thus allowing the outer edges of the ring 19 to come into contact with the shaft 14. As the shaft and the ring rotate, the lubricant is driven upwards from the reservoir 18 through the inner groove .25 24 and two outer grooves 52 and 54 (one on each side of the ring 19). The lubricant is collected and scraped from the groove 24 by the blade 42, after which it is deposited on and against the shaft and the bearing surfaces. The preload of the cantilever blade 42 provides the resilient property minimizing contact between the ring and the shaft and thereby minimizing start-up wear of the elements, while helping to stabilize the ring during operation. high speed operation.

As shown in Figures 2 and 4, the journal and ring normally rotate towards the fixed end of the blade. As a result of the wide configuration of the sheet at the fixed end, the stiffness of this blade increases from the leading edge to the fixed end. This wide structure also serves to collect the scraped lubricant and to direct it into the axial spreading groove (not shown) of the bearing during operation of the ring.

The front end or free end 48 of the blade, as well as its position in the groove 24 exert a chasing effect on the ring, thus preventing excessive lateral friction of the ring in the annular groove 16. Further, this free end provides external damping and the rigidity of the ring following the establishment of a hydrodynamic pressure between the blade and the ring. As the speed of the ring increases with increasing journal speed and as more lubricant is driven upward by the ring, the blade is forced out , roughly in position 50B shown in figure 4. In this way, the outward movement gives rise to a divergent wedge-shaped configuration creating, with the hydrodynamic oil pressure generated, a braking mechanism for the ring by stabilizing the latter during high-speed operation while as the oil distribution increases. In this way, it is no longer necessary to machine different numbers of grooves in the ring for journals having different speeds and sizes. As the journal speed increases more, the divergence effect becomes more pronounced. The ring assumes approximately the position shown at 50C in Figure 4, which is the desired effect, since both the more pronounced divergence exerts an even better stabilizing influence and slower ring speed at higher speeds. 5 upper trunnion. Therefore, stability is inherent in the higher speeds of the ring with an oscillating movement due to the divergent wedge configuration.

Figure 7 gives a diagram of the effects exerted by variations in the depth of throat 24 on lubricant delivery for different shaft speeds. All three rings tested were identical except for variations in the depth of the internal groove which was D = 1.05mm, D = 1.52mm and D = 3.20mm. From these results, an optimum depth of about 1.52 mm was chosen to achieve about twice the oil distribution provided by rings having shallower or deeper grooves. Ring 19 with a groove of about 1.52mm depth was called ring # 5 and was tested with and without the cantilevered oil blade 42.

The results are illustrated by a diagram in figure 6. By carrying out the test on the ring without the blade, the instability was established at an approximate speed of 1,800 revolutions / minute for the journal and at an approximate speed. of 180 revolutions / minute for the ring, while the journal 30 could not operate beyond about 2,500 revolutions / minute. Oil distribution through the ring was limited to approximately 1,200 cc / minute. Tests carried out with the same ring comprising the blade 42 have enabled the journal to operate up to and above 3500 revolutions / minute with an oil distribution of approximately 2,100 cm3 / minute at 1,800. revolutions / minute, while obtaining an oil distribution of about 3,200 cm3 / minute at 1,800 revolutions / minute, the graph of these last results being given in figure 7. The two tests, a diagram of which is given in Figures 6 and 7, were carried out with SAE 20 lubricant by weight. Accordingly, the increased oil distribution seen in Figure 7 can be attributed to a higher lubricant temperature, which, in the test illustrated in Figure 7, was 48.8 ° C at 1 ° C. inlet, while in the test illustrated in Figure 6 the lubricant temperature was 37.8 ° C at the inlet. The effects of variations in lubricant viscosity are shown in a diagram in Figure 9 for lubricants having SAE values ​​of 10, 20 and 30 by weight. As can be seen, heavier lubricants provide marked increases in oil distribution, which is an important and desirable factor, particularly in large bearings where heavier lubricants and speeds are commonly used. higher.

Where the bearing shells have bidirectional capacity, additional collecting means such as a collecting blade 60 is attached to the bearing lining 40 using suitable elements such as screws 62, with the blade 60 being disposed opposite the bearing. cantilever blade 42 illus30 trée in figure 5. The collector blade directs the lubricant distributed in the feed groove of the bearing shaft (not shown) where it is distributed to finally return to the reservoir and be taken up by the grease ring, then recycled. Figure 10 gives a diagram of oil distribution as a function of shaft speed for a two-way bearing shell and for rotation to and away from the cantilever blade 42. Although there is a slight decrease in oil distribution, the distribution is still increased over that obtained with a conventional grease ring alone. Accordingly, it is desirable to incorporate the manifold 60 into a bidirectional bearing and thus, there is possibly no longer a need for an external lubrication system. When this external system is needed, whatever the dimensions of the bearing or other factors, it is always advisable to add the manifold due to the rapid increase in oil distribution that is observed. from the start of operation, thus minimizing start-up wear of the bearing, shaft and ring itself.

Although one embodiment of a bearing lubricating device has been illustrated and described in detail herein, as well as a modification thereof, various other changes and modifications can be contemplated without departing from the scope. of the present invention.