GB1577562A

GB1577562A - Sleeve bearing

Info

Publication number: GB1577562A
Application number: GB21368/77A
Authority: GB
Original assignee: Gould Inc
Current assignee: Gould Inc
Priority date: 1976-05-26
Filing date: 1977-05-20
Publication date: 1980-10-29
Also published as: JPS52145654A; IT1084265B; JPS60182524U; BR7703383A; CA1083211A; DE2723585A1; DE2723585C2

Description

(54) IMPROVED SLEEVE BEARING (71) We, GOULD INC., a corporation organized and existing under the laws of the State of Delaware, of 10 Gould Center, Rolling Meadows, Illinois 60008, United States of America, do hereby declare the invention for which we pray that a patent may be granted to us, and the method by which it is to be performed, to be particularly described in and by the following statement: Background of the Invention This invention pertains to the art of bearings and more particularly to bearings of the type having an oil film or layer disposed between the bearing surface and a journaled member.

The invention is particularly applicable to half-shell bearings used in internal combustion engines in cooperation with crankshafts and connecting rods and will be described with particular reference thereto.

However, it will be appreciated by those skilled in the art that the invention has broader applications and may be used in many other environments of this general type.

As is known, internal combustion engine main and connecting rod bearings are subjected to repetitive loads which can ulti mately fatigue the bearing alloy if it does not have sufficient strength tor the particular applications involved. Metallurgical advances through the years have produced progressively stronger alloys, however, those metallurgical alterations and advances which have thus far been made appear to have reached the limit insofar as improving or increasing fatigue resistance. The primary reason for this is that in order to make an alloy stronger, it must be made harder or alloyed with higher percentages of elements which are not, themselves, good bearing materials.

Bearings constructed from bearing materials which are made stronger by either of the two above noted means will, while theoretically having a higher fatigue strength, have a much greater tendency to seize during operation and will invariably fail from such seizure.

Typical internal combustion engine main and connecting rod bearings are comprised of halfshells or semicircular arrangements constructed from conventional bearing materials and are installed within the engines in a manner well known in the art. Because these particular engine components are received in what amounts to continuously circulated oil bath, there is a clearance area provided between the outer peripheral surface of the journaled member or crankshaft and the inner peripheral surface of the bearing. This arrangement not only facilitates relative rotation between the parts in the proper manner during normal engine operation but more importantly, facilitates the creation of an oil film there between which acts as a load supportirg medium during engine operation. The hydrodynamic wedge effect which causes the shaft or journaled member to float on an oil film is a result of there being a difference between shaft and bearing radii at the associated areas thereof.

For some period of time, it was believed in the industry that fatigue of a bearing alloy could be expressed in terms of maximum unit load (peak load divided by projected bearing area, i.e., length x shaft diameter). However, this belief or concept did not satisfactorily explain why bearings with the same area, but different lengths and diameters, exhibited different fatigue lives under identical loading. Modern analytical methods, specifically the journal orbit analysis, have now explained this phenomenon. Basically, geometric factors such as bearing length, diameter and clearance affect the peak pressure developed in the load supporting oil film. Those bearings exposed to higher pressures fatigue sooner or to a greater extent in the same amount of time as bearings exposed to lower pressures.

Through the use of bearing orbit analysis techniques, it can be shown that the peak oil film pressure developed in a bearing increases in an essentially linear fashion with increased bearing clearance. Thus, a seemingly ready answer for improving fatigue strength for these bearings is to reduce the oil film pressures by simply reducing the bearing clearance. However, a simple reduction in such clearance is not generally preferred since it would reduce the "slop" between the bearings and journaled member and thus be more sensitive to misalignments. Moreover, such a reduction in clearance would also cause the bearing to run hotter that one of normal clearance and it could, therefore, burn up during normal use.

The subject invention contemplates a new structural arrangement which overcomes all of the above referred to problems and others and provides a new bearing structure which is simple in design, provides improved fatigue resistance over conventional bearing structures and which is readily adaptable to use in a number of environments.

Erief Description of the Invention In accordance with the present invention, there is provided a bearing assembly comprising a bearing which surrounds a substantially cylindrical journaled member which is circumferentially differentially loaded during an operational cycle and is separated from the bearing by a lubricant film, the bearing comprising a portion which has a lesser inner radius than the mean inner radius of the bearing and confronts the more highly loaded portion of the journaled member, and another portion which has a greater inner radius than the mean inner radius of the bearing and confronts the lesser loaded portion of the journaled member, so that during operation the more highly loaded portion of the journaled member is disposed more closely to the bearing.

By means of this structure, the first portion of the bearing will " see " a journaled member radius more nearly equal its own radius during the high load portion of the bearing cycle so that a lower oil film pressure will develop. The second portion of the bearing is disposed in a manner so that it will receive oil film pressure during the low load portion of the bearing cycle.

In accordance with another aspect of the present invention, the portions of the bearing with respectively lesser and greater inner radius have, respectively a greater and a lesser radial thickness.

The bearing assembly may comprise an outer shell, an intermediate layer and an overlay layer, and the differences in the radial thickness of the bearing are provided by differences in the thickness of the intermediate layer.

The bearing may comprise two semicylindrical halves with opposed mating faces, the inner circumferential surfaces may be flush at their partition line, or may be concavely shaped at the parting line so that the opposed mating faces abut along their entire length.

Each of the portions of the bearing may comprise one of the halves of the bearing.

A bearing lining may be provided which is thinner on the portion of the bearing with the lesser inner radius and thicker on the portion of the bearing with the greater inner radius.

Brief Description of the Drawings The invention may take physical form in certain parts and arrangements of parts, a preferred embodiment of which will be described in detail in this specification and illustrated in the accompanying drawings which form a part hereof and wherein: Figure 1 is a perspective view of a halfshell bearing arrangement to which the subject invention is particularly adapted for use; Figure 2 is a front elevational view in partial cross-section showing a conventional prior art halfshell bearing arrangement as installed in a typical internal combustion engine environment; Figure 3 is a view similar to Figure 2 showing the concepts of the subject in Invention as incorporated into a connecting rod bearing for an internal combustion engine; and, Figure 4 is an enlarged partial view of an arrangement similar to that of Figure 3 and showing the concepts of the subject invention as incorporated into a trimetal type of sleeve bearing.

Description of the Preferred Embodiment Referring now to the drawings wherein the showings are for purposes of illustrating the preferred embodiment of the invention only and not for purposes of limiting the same, the Figures show a pair of halfshell bearings A and B of the type normally employed on, for example, internal combustion engine main and connecting rod bearings.

Halfshell bearing A is comprised of a band-like semi-circular body 10 and hnlfshell B is comprised of a band-like semicircular body 12. Body 10 has end faces 14,16 and body 12 has end faces 18,20.

Bodies 10,12 are dimensioned and configured so that end faces 14,18 and 16,20 will substantially mate with each other.

Bodies 10,12 also include inner bearing surfaces or areas 22,24 and outer surfaces 26,28 respectively. These halfshell bearings may, of course, be constructed from conventional bearing materials in any conventional manner. For internal combustion engine usage such as that to which the subject invention is particularly directed, materials such as, but not limited to, lead base babbitt, aluminium alloys and copper-lead alloys are employed for the inner bearing surfaces or areas 22,24 of bodies 10,12.

With reference to Figure 2, description will hereinafter be made to a prior art connecting rod bearing employing the bearing halfshells shown in Figure 1. Here, the crankshaft or journaled member is shown in cross-section and generally designated by numeral 30. The shaft is substantially circular in cross-section and is shown as having a radius rr at the outer peripheral surface thereof.

The connecting rod bearing is comprised of half-shell bearings A and B disposed in a closely spaced relationship with crankshaft or journaled member 30. To provide the desired bearing installation, an upper rod bearing housing 32 which includes con necking rod 34 as an integral part thereof is provided along with a cooperating lower rod bearing housing 36. Each housing has a generally semi-circular opening adapted to closely receive bodies 10,12. Specifically, housing 32 includes a semi circular receiving or mounting surface 38 adapted to closely receive outer bearing surface 26 of body 10 and lower rod bearing housing 36 includes a semi-circular receiving or mounting surface 40 adapted to closely receive outer surface 28 of body 12.

Moreover, housing 32 includes a pair of outwardly extending flanges 42 and housing 36 includes a pair of similar connecting flanges 44 with flanges 42,44 dimensioned and located so as to mate with each other.

The housings are rigidly connected together by means of threaded fasteners 46 in a convention,al manner. Bodies 10,12 are received in housings 32,36 in the manner shown in Figure 2 with substantially mating ends 14,18 and 16,20 in an abutting relationship with each other. Housings 32,36 and bodies 10,12 are dimensioned so that this abutting area is in substantial alignment with the parting line or line of separation between the housings themselves at flanges 42,44. Bodies 10,12 are retained in the housings by the dimensioned relationship between the components when the housings are tightly connected together by r:leans of fasteners 46 and the bodies typically and advantageously include small tabs or lips as shown in Figure 1 for purposes of assuring proper alignment between them. The structure described with reference to Figure 2 is deemed conventional and is well known in the art so that further elaboration on the specifics thereof noted above is deemed unnecessary.

As shown in Figure 2 in an exaggerated form, an annular clearance area a is provided between the outer peripheral surface or crankshaft or journaled member 30 and inner bearing surfaces 22,24 of halfshell bearing bodies 10,12. In practical application, this clearance is fairly small and may only comprise a few thousandths of an inch.

The size of the clearance will vary depending upon the specific application of the bearing within an internal combustion engine as well as the characteristics of the engine design itself. Also in the prior art, both halfthell bearing bodies 10,12 have had a substantially identical thickness designated tin Figure 2 and this thickness may vary fronn a few hundredths of an inch to a tenth of an inch or so. With this substantially constant thickness, a radius of curvature r is defined from the center of the bearing outwardly to bearing surfaces 22,24.

Because of clearance area cr, radius r will be greater than radius rr. In order to facilitate ease of understanding the spacial relationships between the components radii r and rr are shown as being coaxial. In actual practice, however, and due to machining characteristics and tolerances encountered during component manufacture, the centers for the bearing and crankshaft radii may be slightly offset relative to each other involving some eccentric relationship between the components. This relationship does not, however, have any effect on the concepts of the subject invention as described herein.

During operation of an engine which incorporates the prior art connecting rod arrangement shown in Figure 2, oil is received in annular clearance area a forming an oil film. This film has a hydrodynamic wedge effect which causes shaft 30 to float on the oil film during engine operation and bearing cycling. During such cycling, the oil film pressure which has a destructive effect on halfshell bearing bodies 10,12 at bearing surfaces 22,24 is particularly troublesome during the upstroke of the connecting rod and through the cylinder bore. It is during this portion of the cycle, i.e., when crankshaft or journaled member 30 is acting against the oil film disposed between that member and inner bearing surface 22 of halfshell body 10 that the peak or maximum oil film pressures are developed. While there are oil film pressures developed between the crankshaft and inner bearing surface 24 of halfshell body 12 during the downstroke, such pressures are not nearly as great or destructive as those incurred on the upstroke.

Thus, and in a typical internal combustion engine environment, the crankshaft bearings are differentially loaded during each cycle and such loading is dependent upon the position of the bearing during the cycle.

When the rod bearings are subjected to repetitive loads such as those described above, it can ultimately fatigue the bearing alloy if the alloy does not have sufficient strength for the application involved.

Accordingly, it has been desired to improve this fatigue strength in order to yield better overall engine operation.

It has been thought that simple metallurgical improvements to bearing materials would solve this problem. However, in order to make a bearing alloy stronger, it must necessarily be hardencd or be alloyed with higher percentages of elements which are not themselves good bearing materials. When these solutions are employed, the bearings, while theoretically having a higher fatigue strength, also have a much greater tendency to seize during operation.

Knowing that a decrease in clearance area a would act to reduce peak oil film pressures developed during bearing cycling, another apparent solution to the problems would be to simply reduce the clearance area itself. Such a reduction in clearance is not generally preferred or desirable since the elimination of some "slop" within the bearing arrangement will render it far more sensitive to misalignment. Moreover, such a bearing arrangement would run hotter than one of normal clearance and could burn up during normal use.

The subject invention focuses on a solution to the above noted problems as shown in Figure 3. The structural solution is shown as it has been incorporated into a conventional rod bearing arrangement identical to the arrangement shown in Figure 2. For this reason and for ease in appreciating the scope of the invention, like components are identified by like numerals with the inclusion of a primed (') suffix while new components are identified by new numerals or letter designations.

Basically, the concepts of the present invention are directed toward increasing bearing fatigue resistance and still maintaining the same overall diametral bearing clearance as has been conventionally employed in order to prevent seizures or excessive operating temperatures. The desired results are achieved by specific modifications made to hallshell bearing bodies 10, 12. More particularly and in accordance with the present invention, the desired results are obtained by increasing the bearing wall thickness in the heavily loaded of the halfshells by a given amount and by decreasing the bearing wall thickness of the lightly loaded of the halfshells by an equal amount. In this manner, the bearing will " see " or be associated with a crankshaft or journaled member radius more nearly equal to its own radius during the high load portion of the cycle. This arrangement results in a lower peak film pressure being developed there than with conventional bearing structures such as that shown and described with reference to Figure 2.

With reference to Figure 3, bearing body 10' has a thickness w and bearing body 12' has a thickness y. Thickness w is greater than the thickness t of the conventional arrangement shown in Figure 2 and thickness y is less than the thickness t. In order to prevent seizures or excessive temperatures when utiiizing the concepts of the subject invention, the increase in thickness w over thickness t is compensated for by a corresponding identical decrease in thickness y under thickness t. Thus, it is possible to maintain the same overall diametral clearance as is used in the present conventional bearings while still achieving better overall operational results insofar as increasing fatigue resistance.

By increasing the thickness of halfshell bearing body 10' to thickness w, the effective radius of curvature x of body 10' is decreased slightly from radius r shown in Figure 2, although radius x is still greater than radius t ' of crallkshaft or journaled member 30'. In Figure 3, the clearance area between inner bearing surface 22' and the outer peripheral surface of crankshaft 30' is designated b and is smaller than clearance area a of Figure 2. Moreover, and due to the decrease of thickness y in Figure 3, radius of curvature z of halfshell bearing body 12' is increased so that it is greater than the radius of curvature x and the radius of curvature r. Accordingly, a clearance area c is defined between inner bearing surface 24' and the outer peripheral surface of crankshaft 30' and which clearance area is larger than clearance area z7 in Figure 2 as well as clearance area b in Figure 3.

With the structural arrangement disclosed above with reference to Figure 3, during the upstroke of the piston when the greatest peak oil film pressures are developed, the destructive forces normally attendant thereto over long periods of time or operation under heavy loads are decreased due to the decrease in the clearance area b. Again, as bearing clearance increases, the peak oil film pressures developed increase in an essentially linear manner so that conversely, a decrease in bearing clearance causes a decrease in peak oil film pressures. Furthermore, since lower oil film pressures are de developed in the downstroke portion of the cycle, the corresponding increase in clearance area c will not be significant insofar as any destruction of lower halfshell bearing 12' during cycling.

Because of the differences in thickness between halfshell bearing bodies 10', 12', un desirecl oil wiper areas would be created at the parting lines between the two bodies at the substantially mating end faces 14', 18' and lug', 20' thereof. To eliminate these wiper areas, shallow reliefs 50, 52, 54 and 56 are provided across the widths of bodies 10', 12' at the end face areas. The radial depths of these reliefs are calculated so that the halfshell bearing bodies will have the same thickness at the parting lines to thereby eliminate the discontinuities resulting in undesired oil wipers.

The various dimensions and clearances shown in Figure 3 have been exaggerated for purposes of appreciating the scope and intent of the present invention. In actuality and with some typical internal combustion engines, the increase in thickness w of body 10' may only be on the order and magnitude of .001" with a corresponding identical reduction in thickness y of body 12'. While these changes may seem minimal from a structural point of view, the arrangement shown in Figure 3 does provide increased bearing fatigue resistance over the conventional prior art arrangement shown in Figure 2. By way of example and based upon the results of a series of calculations based upon a diesel engine rod bearing, an increase in the upper halfshell bearing wall thickness of .001" with a corresponding reduction in the lower halfshell bearing wall thickness resulted in a reduction in peak oil film pressures of approximately 35 per cent.

While the subject invention has been described with specific reference to adaptation and use on a connecting rod bearing, it should also be appreciated that the invention has been equally adapted for use on the engine main bearings. When adapted for use on main bearings, however, the thicker walled halfshells are employed on the lower main bearing positions with the thinner halfshells being employed in the upper main bearing positions. The subject invention is also adaptable to use in other environments and installations where bearing fatigue resistance is a problem and where an oil film is disposed between the bearing and the journaled member to provide a hydrodynamic wedge effect thus developing peak oil film pressures during cycling which could have a destructive effect on the bearings themselves.

For example and with reference to Figure 4, the invention is readily adaptable to use in heavy duty sleeve bearings more conventionally known as the trimetal type. For ease of illustration and appreciation of this modification, like components are identified by like numerals including a double primed (") suffix and new components are identified by new numerals. The various dimensions and clearances shown in Figure 4 have been conveniently exaggerated for purposes of appreciating the scope and intent of the invention.

In Figure 4 semi-circular bodies 10", 12" are each comprised of three distinct laminations or layers 60, 62, 64 and 70, 72, 74 respectively. Layers 60, 70 comprise steel backings. Intermediate layers 62, 72 comprise a bearing alloy capable of withstanding high loads. These layers are typically comprised of an alloy of copper-lead-tin or aluminum although other alloys could also be advantageously employed. Layers 64, 74 comprise thin electroplated or cast layers of a soft material. Typically, these layers are comprised of lead-tin, lead-tin-copper, tin base babbit or lead-indium materials although other materials could also be utilized.

Layers 64, 74 comprise what are conventionally termed overlay layers and provide the bearing structure with the ability to embed dirt particles circulating in the oil therein. They also furnish "slippery" surfaces to prevent shaft-to-bearing seizure at start-up and shut down and are further such that they may "adjust" to minor irregularities in geometry and / or alignment. For many years the thickness of layers 64, 74 were the same and have been in the nominal range of 0.001" for typical diesel truck engines. As engine loads were increased, however, this thickness had to be reduced to around nominally 0.0006" for preventing premature loss of overlay layers 64, 74 due to fatigue.

Since the fatigue resistance of a bearing alloy decreases as its thickness increases, the above change solved the premature fatigue problem but simultaneously introduced new problems. First, the overlay layers wore out sooner simply because there was less overlay to wear. Second, the overlay could not absorb as much circulating dirt. Once the overlay layers become worn, the harder intermediate layers 62, 72 were exposed to the shaft. The resultant shaft-bearing assembly was then much more prone to seizure and susceptible to damage caused by circulating dirt.

In applying the concepts of the subject invention to a trimetal type sleeve bearing in the same manner described above in sub stantial detail, the total thickness of layers 62, 64 has been made slightly thicker than the nominal thickness and the total thickness of layers 72, 74 is then made thinner by a corresponding amount. Steel backing layers 60, 70 each have the same thickness.

Similar to the Figure 3 arrangement, halfshell or body 10" which includes layers 62, 64 is positioned to receive the heavier loading during use or cycling. Halfshell or body 12" which includes layers 72, 74 is positioned to receive the lighter loading during use or cycling. The increased total thickness of layers 62, 64 provides a smaller radial clearance b" between the surface thereof and the surface of crankshaft or journaled member 30". The reduced total thickness of layers 72, 74 provides a larger radial clearance c" between the surface thereof and the surface of the crankshaft or journaled member.

Insofar as the relative difference in total thickness between layers 62, 64 and 72, 74 is concerned, the thicknesses of layers 64, 74 are generally equal to each other with the thickness of layer 62 being greater than the thickness of layer 72. This is primarily for the reason that it is easier and less costly for practical production reasons to let overlay layers 64, 74 be of the same magnitude.

However, it is possible and sometimes desirable to also vary the thickness of overlay layers 64, 74. In that instance, overlay layer 74 of the thinner halfshell bearing body, i.e., body 12", is made thicker than overlay layer 64 of the thicker halfshell bearing body, i.e., body 10". This modification may be desirable in order to provide better embeddability and is possible because the thinner body 12" is only subjected to light loads so that, therefore, the additional thickness of overlay layer 74 will not suffer fatigue. Here again, however, necessary alterations to the layered halfshell bearing bodies to accommodate the alternative thicknesses for layers 64, 74 will be compensated for in the intermediate layers.

Overall operation of the modified arrangement of Figure 4 is substantially identical to that hereinabove described with reference to Figure 3. However, and specifically as to heavy-duty trimetal type sleeve bearings, utilization of the subject inventive concepts facilitates use of thicker more wear resistant and dirt tolerant overlays. Accordingly, layers 64, 74 may be dimensioned in a manner to thereby extend overlay layer wear life and dirt ingesting capabilities without simultaneously sacrificing fatigue resistance.

The invention has been described with reference to the preferred embodiment.

Obviously, modifications and alterations will occur to others upon the reading and understanding of this specification.

WHA T WE CLAIM IS: 1. A bearing assembly comprising a bearing which surrounds a substantially cylindrical journaled member which is circumferentially differentially loaded during an operational cycle and is separated from the bearing by a lubricant film, the bearing comprising a portion which has a lesser inner radius than the mean inner radius of the bearing and confronts the more highly loaded portion of the journaled member, and another portion which has a greater inner radius than the mean inner radius of the bearing and confronts the lesser loaded portion of the journaled member, so that during operation the more highly loaded portion of the journaled member is disposed more closely to the bearing.

2. A bearing assembly according to claim 1 in which the portions of the bearing with respectively lesser and greater inner radius have, respectively a greater and a lesser radial thickness.

3. A bearing assembly according to claim 1 or claim 2 in which the bearing comprises an outer shell, an intermediate layer and an overlay layer, and the differences in the radial thickness of the bearing are provided by differences in the thickness of the intermediate layer.

4. A bearing assembly according to any preceding claim in which the bearing comprises two semi-cylindrical halves with opposed mating faces.

5. A bearing assembly according to claim 4 in which the inner circumferential surfaces of the halves are flush at their parting line.

6. A bearing assembly according to claim 4 in which the inner circumferential surfaces are concavely shaped at the parting line so that the opposed mating faces abut along their entire length.

7. A bearing assembly according to any of claims 4 to 6 in which one of the halves comprises said portion of the bearing, and the other half comprises said another portion of the bearing.

8. A bearing assembly according to any preceding claim including a bearing lining which is thinner on the portion of the bearing with the le

Claims

**WARNING** start of CLMS field may overlap end of DESC **. are generally equal to each other with the thickness of layer 62 being greater than the thickness of layer 72. This is primarily for the reason that it is easier and less costly for practical production reasons to let overlay layers 64, 74 be of the same magnitude. However, it is possible and sometimes desirable to also vary the thickness of overlay layers 64, 74. In that instance, overlay layer 74 of the thinner halfshell bearing body, i.e., body 12", is made thicker than overlay layer 64 of the thicker halfshell bearing body, i.e., body 10". This modification may be desirable in order to provide better embeddability and is possible because the thinner body 12" is only subjected to light loads so that, therefore, the additional thickness of overlay layer 74 will not suffer fatigue. Here again, however, necessary alterations to the layered halfshell bearing bodies to accommodate the alternative thicknesses for layers 64, 74 will be compensated for in the intermediate layers. Overall operation of the modified arrangement of Figure 4 is substantially identical to that hereinabove described with reference to Figure 3. However, and specifically as to heavy-duty trimetal type sleeve bearings, utilization of the subject inventive concepts facilitates use of thicker more wear resistant and dirt tolerant overlays. Accordingly, layers 64, 74 may be dimensioned in a manner to thereby extend overlay layer wear life and dirt ingesting capabilities without simultaneously sacrificing fatigue resistance. The invention has been described with reference to the preferred embodiment. Obviously, modifications and alterations will occur to others upon the reading and understanding of this specification. WHA T WE CLAIM IS:

1. A bearing assembly comprising a bearing which surrounds a substantially cylindrical journaled member which is circumferentially differentially loaded during an operational cycle and is separated from the bearing by a lubricant film, the bearing comprising a portion which has a lesser inner radius than the mean inner radius of the bearing and confronts the more highly loaded portion of the journaled member, and another portion which has a greater inner radius than the mean inner radius of the bearing and confronts the lesser loaded portion of the journaled member, so that during operation the more highly loaded portion of the journaled member is disposed more closely to the bearing.

8. A bearing assembly according to any preceding claim including a bearing lining which is thinner on the portion of the bearing with the lesser inner radius and thicker on the portion of the bearing with the greater inner radius.

9. A bearing assembly according to any preceding claim in which the bearing comprises layers of different materials.

10. A bearing assembly substantially as hereinbefore described with reference to and as illustrated in any of Figures 1, 3 and 4 of the accompanying drawings.