DE3036418A1 - VIBRATION DAMPER - Google Patents

Description

Pater.ta.iwäite Dipl.-Ing. Curt Wallach

r. Dipl.-Ing. Günther Koch

Dipl.-Phys. Dr Tino Haibach

Dipl.-Ing. Rainer Feldkamp

D-8000 Munich 2 Kaufingerstraße 8 Telephone (0 89) 24 0275 Telex 5 29 513 wakai d

_Date: September 26, 1930

Our reference: l6 956 - K / Ap

Applicant: Barry Wright Corporation

Pleasant Street
Watertown, Massachusetts
United States

Title: Vibration Damper

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3036419

The invention relates to vibration dampers that used are used to connect two components of a structure in such a way that a transmission of vibrations and noise is reduced between the two components. In particular The invention relates to vibration dampers in the form of elastic elastomer bushing between the components are inserted and through which a fastening bolt is passed.

Vibration damper in the form of elastic sleeves made of rubber or the like, which are inserted into an opening of a component and through which a fastening bolt is passed in order to cause the components to be braced against one another, are known in numerous embodiments. In general, such vibration dampers are in the form of two thick-walled ones hollow cylindrical tubes each with a radially extending flange at one end, or they consist of one Combination of such a flanged tube and an elastic washer. The pipes are dimensioned so that they fit sealingly into a recess that runs through the component to be fixed and through the in turn, a screw bolt can be passed through with a snug fit. The flanges (or the flange and washer) are designed so that they attack the surface of the adjacent component and enclose the opening, and they are held in mutual engagement by the bolt or other tensioning means.

As is known, such isolators act by elastic connection of the two components to a mechanical oscillator to form its natural frequency as a result of the training is less than any vibration frequency that it is desired to dampen by at least about $ 29. The amount of damping depends

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on the ratio of the two frequencies and it becomes greater the greater the ratio of the frequency to be attenuated becomes a natural frequency. The natural frequency of the transition oscillation of a component of a given mass that exceeds a resilient isolator is supported is directly proportional to the square root of the spring stiffness of the Isolator, and this in turn depends on the type of stress to which the isolator is subjected and on the Composition of the insulator as well as its geometry and dimensions. In the case of bushing insulators, the in the grommet occurring loads recognized as compressive loads, and the spring stiffness of such Isolator is the product of the compressive modulus of the material and the cross-sectional area of the isolator normal to that worn Load divided by the thickness of the insulator in the direction of the applied force. For elastic rubbery fabrics is the modulus itself depends on the shape, since such materials are essentially incompressible and their elasticity obtained by elastic flow, the free surfaces of the elastomeric material being bulged outward when a load is applied. For a given load, the modulus increases when the cross-sectional area that supports the load increases and when the free surface (bulge surface) decreases. The module's dependence on these two areas is usually called a dependency on one expressed in individual parameters, namely the shape factor, which is defined as the ratio of the load-bearing area to the Bulge area. The larger the design factor (form factor), the larger the module becomes.

The stiffness and, accordingly, the vibration damping that is achieved by such an isolator can be in various Directions be different. A fundamental one

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The problem with the formation of such a vibration damper is the required rigidity in various Provide directions with the vibration damper remaining within a predetermined external shape and size. at Bushing isolators is the isolator or vibration damper generally symmetrical about an axis, and therefore only needs to consider the axial and radial stiffness to be four. The axial loads are absorbed by compression of the flange head of the isolator, while radial loads can be recognized by the smaller diameter in the middle.

Attenuating low frequencies is difficult, especially as far as the axial load is concerned. As mentioned, the natural frequency changes with the square root the spring stiffness, and accordingly results in a 10-fold magnification the spring stiffness only leads to a 3-fold increase in the natural frequency. In known vibration dampers, the Spring stiffness in the axial direction is typically reduced by reduced load bearing areas, which is reduced by reducing the radius of the flange head, or by increasing the thickness of the flange head, or both Measures. With different spring stiffness of the insulator by changing the flange radius for an upright shark tion of a constant flange thickness results when the load-bearing surfaces change as the square of the radius that the Shape factor only changes to the same extent as the radius, since the bulge area changes directly with the radius. In order to obtain a higher order dependency of the form factor on the radius, in order to keep the spring stiffness within wider limits To be able to change, you have to increase the thickness of the flange as the radius decreases. To a desired low to maintain axial spring stiffness by such a design, one often approaches inexpedient designs insofar as

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the required flange diameter is too small to fix the vibration damper properly on a component, or the flange becomes too thick, which is the permissible Space is exceeded, or a less stable configuration results.

A modified constructive approach is to make radial grooves in the flange, creating the load bearing surface and the form factor of a flange that fills a given cavity can be varied. It however, it should be noted that for a given number of grooves of a given depth, the shape factor is again not so changes rapidly like the load-bearing area since some change leads to a change in the load-bearing area in the same sense as the bulge area. The construction analysis such a flange requires a greater laboratory effort than a simple flange, and moreover are the tools required to make such a part are much more complex and expensive.

The invention is therefore based on the object of a bushing insulator To create (vibration damper), which can be designed so that it has a low axial spring stiffness without either the flange diameter becoming too small to properly attach it to the component or the flange becomes too thick, thereby exceeding the available space.

Another object of the invention is to provide a family of such vibration dampers in which the flange heads of the isolators all have the same spatial extent, while attaining different axial spring stiffnesses can be.

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The invention also aims to provide a training of vibration dampers that ensure easy design analysis and simple tools.

The invention also aims to create a vibration damper, the maximum stability for a given load carrier area having.

The task at hand is achieved by a two-part construction of the vibration damper, with one part of one elastic disc is formed, which is provided with a concentric inner recess and a centric circular Has opening, wherein at least one of the discs flange-like at one end of an elastic cylindrical tube is fixed. The axial spring stiffness of the vibration damper can be changed by maintaining the external dimensions of the elastic disc, the diameter the recess in the disc is changed. By change the spring stiffness, the natural frequency of the vibration damper can be reduced, the outer diameter of the Flange is maintained, or it can be a small form factor for a given loading area and flange thickness can be obtained because the bulge area actually increases as the load bearing surface drops.

Grommets made according to the invention can therefore have different axial spring stiffnesses, although accommodation is still possible under the same spatial conditions. In particular, this type of Durable guide sleeves with minimal axial spring stiffness with a given flange outer diameter and axial thickness.

If the sleeve is in the form of concentric surfaces of revolution,

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who are cut in their normal planes, then is there? Tooling for molding the elastomer is relatively simple and can be easily manufactured at low cost, in particular compared to radially grooved sleeves. This shape itself leads to a simpler mathematical one Analysis for design purposes than was the case with radially grooved sleeves.

Since the axial loads are circumferentially supported by a flange of maximum permissible diameter and minimum thickness and not over a number of radially directed cushions or over a smaller diameter or an axially thicker one Flange are distributed, the support sleeve according to the invention is relatively more stable than known bushings.

Exemplary embodiments of the invention are described below with reference to the drawing. In the drawing show:

1 is an isometric view of one end of a bearing sleeve constructed according to the invention;

FIG. 2 is an isometric view of the other end of the bearing sleeve according to FIG. 1;

FIG. 5 shows an axial section of the bearing sleeve according to FIGS. 1 and 2;

4 shows an axial sectional view of two bearing sleeves of the type according to FIG. 1, assembled to form a shock absorber between two components;

5 shows an axial section of a modified embodiment.

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In Figures 1 to 3, a bearing body 20 is produced according to the invention, and this has a sleeve 22 with an outer splash 24 attached to one end of the sleeve is molded. The sleeve 22 and the saddle 24 are preferred made in one piece, and this bearing body can be made of synthetic or natural rubber, for example neoprene or polyurethane rubber. A preferred material is fluorosilicone rubber.

The sleeve 22 is in the form of a thick-walled cylindrical tube with an outer surface 26 and a concentric one inner surface 28, as can be seen from FIG. The end of the sleeve 22, which faces away from the flange 24, is formed into a flat end ring surface 30. In the The sleeve is a tube 32 adjacent to the inner surface 28 inserted. The tube 32 has the shape of a circular cylinder with an axial bore 34. The axial length of the tube 32 is somewhat greater than the axial length of the sleeve 22. As the length of the sleeve 22, the distance between its end ring surface 30 and the ring surface 40 of the flange 24 understood. The tube 32 is inserted into the sleeve 22 that it is a little Dimension protrudes beyond the end ring surface 30, as in detail will be described below. The tube 32 is adjacent to the end ring surface with smooth end ring surfaces 36 and 28, respectively 30 or on which the end ring surface 30 is worn End. The tube 32 is made of rigid material and according to the preferred embodiment, it consists of aluminum. However, other materials could also be used find, provided that they have the required rigidity. According to a preferred manufacturing method the sleeve 22 is formed over the tube 32 so that an integral unit is created. It is clear, however, that the sleeve is also made separately with a bore which forms the inner surface 28, the

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Pipe is then inserted later and connected with a suitable adhesive.

The flange 24 projects outwardly from each end of the sleeve 22 which faces away from the end face 30. Of the Flange consists of an annular disc extension of the sleeve, which runs concentrically to the sleeve. The flange 24 is with two substantially smooth planar smooth surfaces 40 and 42 adjacent to and facing away from outer surface 26 of sleeve 22, respectively. The flange is dimensioned so that the flange surface 42 is at the same distance or further from the end face 30, as that Surface 38 of tube 32. The flange has an outer cylindrical surface 44, the surface 42, which is not continuous over the radial extent of the flange 24 runs, has the shape of an annular lip which is provided with an annular recess 45 which is concentric to the outer cylindrical surface 44 and defined by inner cylindrical surface 46 and bottom 48. The axial extent of the inner cylindrical surface 46 is selected so that it is smaller than the axial extent of the cylindrical surface 44. According to the invention, the axial dimension (length) and diameter of the inner cylindrical Dimension surface 46 taking into account the required axial spring stiffness, as described below is described. The central area of the recess 45 becomes typically occupied by tube 32. A filler ring 49 * which, however, is not essential for the invention, can be formed between bottom surface 48 and tube 32 if required.

Two such bearing sleeves 20, which are designed according to the teachings of the invention, are used around one Form vibration damper. From Figure 4 is in a section

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can be seen that two components 50 and 52 over the invention trained vibration damper are connected, which has two bearing sleeves 20a and 20b, which in conjunction with a washer and a bolt 56 can be used. According to the preferred embodiment, the bearing bushes 20a and 20b are identical in size and composition, but they could also be different from one another in terms of their structural shape differ. For example, the axial length of the sleeve 22 and tube 32 of a bearing can be shortened to the point where where one of the sleeves that form the vibration damper only forms an elastic washer that makes one rigid Has hollow core, the sleeve portion of the insulator is formed entirely by the other camp. The bolt 56 passed through the washer 54 and the tubes 32 and is screwed into a threaded hole of the component 52, holds the two bearing sleeves 20a and 20b in a hole of the component 50 fixed, which is dimensioned so that it receives the sleeves P.2. The flange 24a of the bearing sleeve is pressurized by the bolt 56 between the washer 54 and the component 50 while the flange 24b of the bearing 20b is held under pressure between the components 50 and 52 in a similar manner will. The amount of compressive force that can be exerted by the bolt is limited, according to the preferred embodiment, by the axial lengths of the tubes 32, which are dimensioned to have a combined length equal to the amount of compression desired is less than the combined thickness of component 50 and unloaded Flanges 24a and 24b. Equal control of the degree of compression of the flanges can be achieved by other means, for example, can be achieved using a torque wrench. In addition, the washer 54 can with an epoxy resin or buckets of other adhesive on the surface 42 of the flange 24a can be set. Instead of this

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the plan 24a is the same as plan 24b and structure and Puncture of the parts of the bearing bushes 20a and 20b, which are provided with the same reference numerals, are the same as in FIG Bearing bush 20 »

As mentioned, a vibration damper can also be produced in this way be that a single bearing sleeve 20 is combined with an elastic washer. An elastic one Washer 20c suitable for such applications is shown in FIG. 5. The washer 20c consists of one Flange 24c and a tube 32c. The tube 32c is sized to be somewhat shorter in terms of axial dimensions than the flange 24c, and it is provided with a flat end ring surface 36c which is in the same plane as the ring surface 40c of the flange is located. The ring surface 40c projects radially from the tube 32c toward the outer surface 44. Otherwise, the flange 24c corresponds to the flange 24, and the pipe 32c corresponds to the pipe 32.

In use, the resilient washer 20c can be combined with a sleeve 20 to form one not shown To form a vibration damper which is constructed in essentially the same way as a combination of two Bearing sleeves 20. For example, the elastic washer 20c can be substituted for one of the bearing sleeves 20 in the construction 4 can be used with the washer so arranged is that its surface 40c touches the component 50 and tube 32c is concentric with the hole through that member. The construction is then done in the same way secured like bearings 20a and 20b. Since the elastic washer 20c is not provided with a sleeve, all radial loads on such a vibration damper

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be exercised, primarily through the sleeve 22 of the bearing assembly 20, which is used in conjunction with the elastic washer. In every other way it is Behavior of the vibration damper, the combination of a bearing sleeve 20 and an elastic washer 20c is formed, like that of a vibration damper, which is composed of two bearing bushes, with the same parts Perform the same punctures on the elastic washer and the bearing sleeve. Accordingly, axial loads can both through the flange 24 of the sleeve as well as through the face 24c the elastic washer.

As far as a vibration damper is formed from two bearing sleeves 20 or one bearing sleeve 20 and a washer 20c they do not differ significantly from one another in the details of construction, assembly and mode of operation, so that the following detailed description refers to a vibration damper which consists of two bearing sleeves.

According to FIG. 4, the movements of the components 50 and 52 absorbed relative to each other by the bearing sleeves 20a and 20b as dynamic compressive loads. The movements of the two components towards or away from one another result in axial compressive loads on the flanges 24b and 24a, while Movements in the direction perpendicular to this axis result in radial compressive forces on the sleeve 22. At a given spatial envelope, i.e. given the outer diameters of the flanges 24 and 24b and the sleeves 22 and given combined axial length of the tubes 32 and thickness of the component 50, the spring stiffness of the insulator depends on the material of the structure and the other dimensional parameters.

The radial spring stiffness depends, among other things, on the axial

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Dimension of the sleeves 22, which together with the already mentioned outer diameter of the sleeves determines the load bearing surface. In addition, the radial spring stiffness depends on the inner diameter of the sleeves, i.e. from the outer diameter of the tubes 22, and this outer diameter of the tubes 32 determines the radial thickness of the sleeves and their bulge areas with the outer diameter of the sleeves. Both Flanges 24a and 24b, the axial and radial dimensions of the outer cylindrical surfaces 44 are predetermined by the desired spatial envelope, and the axial and radial dimensions of the inner cylindrical surface 46 can be changed to include the area of surface 42, i.e. the load bearing area for axial loads and the internal cylindrical surface 46 to change the bulge area to part. Bearing sleeves made according to the invention can therefore have different axial and radial Peder stiffnesses, nevertheless the outer outline may be similar. In particular, this type of bearing sleeve can have a minimal axial spring stiffness have a given flange outer diameter and axial thickness.

As far as the shape of the bearing sleeves represents a series of concentric surfaces of revolution that cut through their normal planes the tool used to shape the elastomeric body is relatively simple. It can easy to be made relatively cheap, especially in comparison with radially grooved bearing sleeves. In this context It is important to note that essentially the same shape can be used to make bearing sleeves with to create identical external dimensions but different spring stiffness, since the dimensions of the cavity 45 can be changed by adding different cores to the shape. As a result, a variety of

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Vibration dampers are produced, which are in their Differentiate between spring stiffness, with only nominally higher plant costs being required than for the production of Vibration dampers with identical spring stiffnesses.

The shape of the bearing sleeves according to the invention also leads even to a simpler mathematical analysis for the design sz than is the case with radially grooved bearing sleeves is possible. Since axial loads also circumferentially through a flange of maximum permissible diameter and minimum axial thickness are supported and not over a number of radially extending cushions or over a smaller diameter or an axially thicker flange are distributed, the bearing structure is relatively more stable than with the known sleeves.

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Claims (12)

Claims l.j vibration damper in the form of an elastic Sleeve for passing through a fastening bolt which braces two components against each other, wherein the vibrations that occur are damped, characterized in that that a sleeve made of elastic material is provided which has a thick-walled hollow cylindrical Tube section has that a flange made of elastic material in the form of a disc concentrically on the outside is attached to the first end of the sleeve, and that a lip of elastic material in the form a ring set circumferentially on the flange that is spreading away from the second end. 2. Vibration damper according to claim 1, characterized in that a rigid, cylindrical tube is firmly inserted into the Axial bore of the elastic sleeve is used. 3. Vibration damper according to claim 2, characterized in that the rigid cylindrical tube is dimensioned and is arranged to protrude from the bore of the sleeve at at least one end by an amount which is greater than the amount by which the axial lip protrudes from the flange at the first end. 130016/080? 303641a 4. Vibration damper according to claim 1, characterized in that that a rigid washer is fixed concentrically with the sleeve on the lip. 5. Vibration damper according to claims 1 or 2, characterized in that that the sleeve and flange are integrally molded with one another and made of the same material are. 6. Vibration damper according to one of claims 1 to 5, characterized in that it is a circular disk made of elastic material having a central circular opening, and that a lip made of elastic material in shape a ring is attached circumferentially on one side of the disc. 7. Vibration damper according to claim 6, characterized in that that a rigid cylindrical tube is tightly inserted into the central opening. 8. Vibration damper according to claim 7, characterized in that the rigid cylindrical tube is dimensioned and arranged to emerge from the circular opening protrudes from the disc by a distance which is no greater than the axial extent of the lip across the disc. 130018/0 9. Vibration damper according to claim 6, characterized in that that a rigid cylindrical washer concentric with the elastic washer the lip is set. 10. Vibration damper according to claim 6, characterized in that the disc and the lip are integral with one another are made of the same material. 11. Vibration damper according to one of claims 1 to 10, characterized in that that it consists of two sleeve parts, namely a disc made of elastic material with a central one Central opening and with a lip made of elastic material in the form of a ring that protrudes on one side of the disc and a sleeve made of an elastic material in Shape of a thick-walled hollow cylinder (or two such hollow cylinders without a disk), where carries a protruding lip on the cylindrical tube at one end. 12. Vibration damper according to claim 11, characterized in that that the two parts are arranged so that the smooth sides of the discs closer together lie than the lipped sides of these disks. 130016/0802 303641 Ij5. Vibration damper according to Claim 12, characterized in that the discs and sleeves consist of the same material. 13ÖÖ16 / 08Ö