PRE-PUSHED BRAKE GASKET
FIELD OF THE INVENTION This invention relates to the design of a fastened gasket, such as a gasket fixed by bolts, and in particular to a gasket that prepares a fastener to give rise to a more uniform tension in the fastener in a gasket load. maximum application. BACKGROUND OF THE INVENTION Bending stress is induced through the rod 11 of a fastener such as a pin in the plane of flexure when the joint is not symmetrical and when the load of the fastener is not in the centerline of the fastener. Figures 1A and IB schematically illustrate two fasteners with the same axial application load (i.e., the load on the shaft of the bolt shank). Figure 1A illustrates a flexure stress superimposed on the plane of the paper (in the plane of flexure, to make the stem of the left convex fastener 11) in the axial pre-tension to subject the stem 11 of the fastener 10 to bending loads and Axial Figure IB illustrates the shank 11 of the fastener 10 subjected only to an axial load (no bending load) equal in magnitude to the axial load 12 on the central axis of the bolt in Figure 1A so that the average stress in both fasteners is the same.
REF.172523 The flexure tension in the fastener in Figure 1A reduces the load carrying capacity of the fastener and gasket. One side 14 of the shank 11 of the fastener in the bending plane has a higher tension than the other side 16, due to the induced bending. This is not a desirable condition because the distribution of tension across the shank of the fastener causes high stress on the side 14 of the fastener shank. A more desirable tension condition at the maximum load would be to have a uniform voltage distribution across the rod of the fastener 11 in the flexure plane under maximum load conditions as illustrated in Figure IB, where the voltages 12, 14 and 16 They are substantially the same. In some cases, the clamped joints can not be designed to eliminate bending stresses in the fastener under all conditions, such as in a connecting rod where the application load is dynamic and therefore changes. The load that is carried by the fastener is related to the average tension in the fastener. In Figures 1A and IB, both fasteners 10 have the same average tension 12 but the fastener in Figure 1? has a maximum maximum tension 14, as a result of the bending stress. If a fault occurs, it would occur at the point along the side with the highest tension 14. Thus, the bending stress added to the axial tension reduces the carrying capacity of a fastener compared to a fastener subjected to the same average tension but with a uniform voltage distribution. With reference to Figure 2A, when a gasket is clamped, an initial axial pre-tension is applied as a result of the apprehension or tension of the fastener. This is represented by the uniform pre-tension components 18. If the joint is not symmetrical, it will be compressed more on one side than on the other side of the fastener hole. This causes the rod of the fastener 11 to be subjected to bending stresses and the load to be applied in a non-uniform manner through the rod of the fastener 11. This is represented by the non-uniform components 20. In addition, if the application load is applied eccentrically to the centerline of the fastener, additional bending will occur in the fastener. The tension components 12, 14 and 16 in Figure 1A are the sum of the uniform components 18 and the non-uniform components 20 in a maximum application load. The design of the fastener joint is limited by the highest tension level in the fastener, including the bending tension, which makes a uniform tension profile, as illustrated in Figure IB, a more desirable option.
The word "fastener" used herein is any type of fastener that has a stem that is subjected to tension forces when applied to a joint, such as bolts, rivets, rods (threaded, fixed, welded, etc.), screws, etc. The word "bending stress" refers to a non-uniform tension across the fastener stem. This invention includes the use of a nut in the joint system, which could act such as a bolt head and therefore the "head" of a fastener includes a nut, a bolt head or screw head, rivet head or rivet tab, etc. BRIEF DESCRIPTION OF THE INVENTION This invention provides a bolt joint that under maximum load conditions in the service application of the joint the maximum tension will be reduced through the bolt shank. This is done by the joint which induces a bending stress on the fastener shank in the plane of flexion of the applied bending stress when the fastener is mounted on the joint. The bending stress induced by the joint is substantially inversely proportional to the bending stress induced in the plane of bending by the maximum application load such that the rod of the clamp is put into service to reduce the maximum stress when the clamp is applied. maximum application load. By the above, the invention also reduces the cyclic main tension to which the fastener stem is subjected. This is especially useful for increasing the fatigue resistance of the fastener. In a useful aspect of the invention, the bending stress induced by the joint is of a magnitude and direction such that it produces a substantially uniform stress distribution across the clamp rod in the bending plane when the maximum application load is applied. , to obtain the full advantage of the invention. In one embodiment of the invention, the gasket has a seat that the fastener bears against to induce tension in the stem, and the seat is biased at a different angle of 90 degrees in an axis of a fastener hole in the parts to through which the shank extends. The seat is oriented in one direction to induce the bending stresses in the fastener, opposite in direction to the bending stresses induced by the maximum application load. Thus, the bending stresses induced by the joint eliminate the bending stresses induced by the application load to reduce the maximum application load on the fastener stem and to reduce the cyclic head tension to which the stem of the fastener is subjected. attacker In another embodiment of the invention, the gasket has sealing faces that face each other and are held together by the fastener, a portion of the gasket faces defines between them an unsupported groove that induces bending stresses in the fastener stem. , opposite in direction to the bending stresses induced by the maximum application load. In other embodiments of the invention, an orifice extending in the parts and receiving the rod of the fastener has a first portion in one of the parts and a second portion in the other part, with the first portion inclined in relation to the second portion. to induce bending stresses in the fastener, opposite in the direction of the bending stresses induced by the maximum application load. These different embodiments of the invention can be practiced alone or in any combination with each other. In a particularly useful form, the gasket is a joint in a connecting rod that connects a support bushing with a rod portion of the rod. A support bushing joint is an especially useful application of the invention because the rod of the fastener is subjected to a cyclic bending stress by the cyclic movement of the connecting rod, such that the pre-tension of the fasteners using the invention , can reduce the maximum application voltage and the cyclic main tension in the bra studs. The foregoing and other objects and advantages of the invention will appear in the detailed description that follows. In the description, reference is made to the accompanying Figures which illustrate a preferred embodiment of the invention. BRIEF DESCRIPTION OF THE FIGURES Figure 1A is a typical bolt tension distribution diagram of the prior art illustrating the stress distribution in a bolt subjected to bending and axial load; Figure IB is a typical prior art bolt tension distribution diagram illustrating the stress distribution in a bolt subjected only to axial load, with the axial load magnitude equal to the load on the bolt axis in the Figure 1A; Figure 2A is a view similar to Figure 1A with the voltage diagram illustrating the total voltage components such as the pre-tension and maximum tension; Figure 2B is a view of a fastener comparable to Figure 2A, but with a pretension and maximum tension distribution produced by a gasket embodying the invention;
Figure 3 is a view of a connecting rod support socket with an angled bolt seat according to the invention, the angle being exaggerated for illustrative purposes; Figure 4 is a view such as Figure 3 but of a typical prior art rod support strut joint; Figure 5 is a view of a rod support flange gasket with the angled gasket faces according to the invention, the angles being exaggerated for illustrative purposes; and Figure 6 is a view of a rod support flange gasket with threaded fastener holes in the rod that is internally angled according to the invention, the angles being exaggerated for illustrative purposes. DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT Referring to Figure 2B, the present invention provides a secured fastener gasket design that provides a substantially uniform stress load distribution on the fastener stem 11 under maximum application load conditions. In Figure 2B, the stress graph shows the initial pre-tension with a lower set of vector arrows 22, 24 and 26 and the application tension by the upper set of vector arrows 28, 30 and 32. Both in the Figure 2A and Figure 2B · the average pre-tension and the average maximum voltage are equal, so each case would be handling the same system load; however, using the invention results in a lower maximum stress under the application load. In the case where the joint application loads are cyclical such as in a connecting rod bearing joint, the main cyclic tension would also be lower. The total cyclic voltage oscillation would remain the same. The schematic voltage graphs of Figures 2A and 2B are simplified in that they do not show any pretension of bending of incidental or accidental joint. If there were pre-tension of joint bending, the horizontal set of pre-tension vectors in Figure 2A would not be uniform (at some angle) and the corresponding pretension vectors in Figure 2B would need to be adjusted to compensate the pre-tension of flexion. The uniform voltage distribution in the maximum application load can be achieved in any number of ways. Currently, typical rod holder bushings are made as illustrated in Figure 3, with each bolt seal seat 35 facing 90 ° to the corresponding bolt hole 37 and the threaded hole 39 of the centerline 38, the The non-threaded hole 37 is in the support socket 42 and the threaded hole 39 is in the connecting rod body 44. This produces a stress distribution as substantially as in Figure 2 ?, with the vectors 18 representing the static tension and vectors 20 representing the dynamic application load. Note that in this case, the maximum tension occurs on the inner side (towards the crankshaft bore of both seats 36). One embodiment of the present invention would be to bias each joint bolt seat 36 in the bolt hole 37 and in the threaded hole 39 of the centerline 38 by some small amount, chosen based on the maximum application load that will be eliminated or compensated . Generally, the angle will be less than one degree, for example, .125 degrees, depending on the magnitude of the application load. The angle must also be in the correct direction, so that it eliminates the bending stress in the maximum application load (dynamic) condition, which is induced by the joint and the application load. This is illustrated in Figure 3. Both seats 36, which are flat according to the illustrated, are machined or formed to make an angle or are biased internally in the direction of the bending plane, to induce bending stresses in each bolt 10. , which are opposite to the bending stresses induced by the application load. In other words, the bolts 10 tend to bow outwardly (convex-external in relation to the axis of the main interior surface 40) in the plane of the paper as a result of the slanted seats 36, while the application load tends to arch the bolts 10 internally ( convex-internal in relation to the axis of the main interior surface 40). The magnitude and direction of the angle of the seats 36 is chosen, and also the torque is chosen at which the bolts 10 are tightened, to produce a substantially uniform stress distribution in the rod of the fastener 10 at the maximum application load , as illustrated in Figure 2B. If in Figure 4, the bolt hole 37, 39 and the bolt joint seat are machined along the same central line of the spindle 38, the centerline of the seat and the bolt will be at 90 degrees to each other under of the manufacturing process, such as the typical gasket shown in Figure 3. An additional or different process is necessary to create the obliquity of the required bolt seat 36. This could be done in many ways. For example, the obliquity of the bolt seat 36 of Figure 4 could be forged in the support bushing 42. Another way would be to machine the bolt hole with an axle along the spindle 38 and to machine the bolt seat with another spindle at a small angle for the orifice drilling spindle. Yet another way would be to create the angle of the seats 35 using the powder metallurgy process to form the obliquity of each bolt seat 36 in the support bushing 42. Another way to create a uniform tension through the bolt rod 11 in the plane of bending at a maximum application load is to make the joint faces, where they face each other near the center of the main interior surface 40, at a small angle for the outer taper to each other to create a groove without small support 48 between each set of gasket faces in the area adjacent to the inner surface 40. This is illustrated in Figure 5. One or both facing surfaces could be angled to create the slit 48. This small angle (greatly exaggerated in the Figure 5; can be less than one degree depending on the magnitude of the application load that will be eliminated) could be machined on the faces, formed by forging or by powder metallurgy, or the joint could be deformed plastically to create the slit, whose last method it could be incorporated in another production process of division by fracture typical of a rod and bushing of a connecting rod. This allows the bushing 42 to flex towards the rod member 44 in the areas of the grooves formed by the angles, which has the effect of subjecting the rods 11 of the fasteners 10 to bending stresses to arch them outwardly. When the bolts 10 are tightened, the slit 48 can be closed or substantially closed, or not. The size of the slots 48 and the torsional force at which the bolts 10 are tightened, are chosen to produce a substantially uniform stress distribution on the shank 11 of the fastener 10 in the flexure plane at the maximum application load, as it is illustrated in Figure 2B. Yet another way to create a uniform stress in the plane of flexure through the pin shank 11 at a maximum application load, would be to create the center line 38A of the threaded hole 39 at a small angle for the bolt hole 37 and the line central 38B (unfolded) of the bolt 10 as illustrated in Figure 6. Once again, the angles of the axes 38A are greatly exaggerated and may be less than one degree in relation to the axes 38B. These externally arch the rods 11 of the bolts 10, as in the previously described modes, to produce a uniform stress distribution through the bolt shank in the flexure plane at the maximum application load, with a reduced cyclic head tension and the reduced maximum stress on the pin shank 11. The angles of the axes 38A and the torsional force at which the bolts 10 are tightened, are chosen to produce a substantially uniform stress distribution in the plane of flexure in the shank 11 of the fastener 10 to the maximum application load, as illustrated in Figure 2B.