CA2181619A1 - Bearing - Google Patents

Abstract

The object of the invention is a bearing, in particular an engine bearing, to support in damping manner masses vibrating in different modes. The bearing is composed of a load connector inserted in fused manner by means of a rubber-metal compound into an elastomer radial spring, said connector being coupled not only to the said radial spring receiving the radial vibration components but also and independently from said radial spring to an axial spring receiving both the static and the axial components of the dynamic load.
The two springs, both the radial and the axial springs, are supported on a bearing housing acting as the matching rest. The axial spring is located inside a hydraulic working chamber formed inside the bearing housing and bounded by the radial spring. The radial spring is designed as a fluid-tight bellows spring forming a second hydraulic or pneumatic working chamber. Preferably the working fluid of this second inner working chamber can be impressed, through a duct and a pressure generator comprising an appropriate control, with a dynamic force envelope which opposes in cancelling manner the dynamic loads applied to the bearing. Preferably the pressure generator is a membrane generator or a solid-body vibrator. The engine bearing of the invention evinces active wide-band decoupling and wide-band damping while offering much improved compactness over the related active conventional hydraulic bearings.

Description

. ~ 21~1~19 BEARING

The invention concerns a bearing damping the vibrating masses it supports, and applicable in particular in the automotive industry as engine bearings to snpport engines md drive units.
As a rule such bearings are elastomer-metal rnmrol-nfl~ amd typically are composed on one hand of a load connector and a matching-rest typicaUy in the form of the bearing housing and on the other hand of a resilient rubber body inserted between the two above L~ Damper-bearings of this kind lvl~du~aLItly are designed as so-called hydraulic bearings wherein a hydraulic operating fluid is so enclosed inside a wL~rking chamber that when the bearing is subjected to a dynamic load which must be damped, the fluid shall be able to overflow tbrough a throttling overflow duct into an escape chamber.
When the load is removed from the bearing~ the operating fluid can flow back into the working charnber. At least part of the walls of the working chamber is constituted by the f ~ f bearing sprmg which usually is more or less conical.
Bearings of this species entail the problem of lacking wide-band efficacy relating to decoupling and damping the load connector and the matching rest when subjected to dynarLuc.impacts and vibrations from a wide number of sources that, especially in auto-= ~
motive applications, will ~ v ~ly act on the bearing.
Accordingly and illustratively a bearing expert merely faces a routine problemif asked to so design a bearing that the low-frequency engine vibrations which are typically between 5 and 10 Hz shall be optimally damped in the fluid bearing. Such an "optimized"
bearing howeYer will trarlsmit body acoustic vibrations in the typically critical low-frequency rarlge of about 100 to 300 Hz as undamped from the load connector to the matching rest or " ~ 21~1619 from the matching rest adapter to the ~ tomr~bil~ body, and so it does for highenergy, ride-caused impacts.
Thousands of CA~ i toward improving damping amd decoupling relating to the load connector and tbe matching rest of fluid bearings of the above kind haYe been carried out. This fact alone may be corlsidered reliable indication that a basic solution to wide-band decoupling a d wide-band dampmg between the load connector and tbe matching rest of a hydraulic beanng is still far away.
Accordingly the present objective invention is based on the tecnnical problem to create a hydranlic bearing for a vibrating mass to be damped, in particular an engine bearing for the automotive industry, which shall eAtensively prevent dynar~uc loads acting on it, dependirlg on the Aind of load and on frequency and amplitude also, from passmg from one of tbe two bearing cormectors to the particuLar other one.
For that purpose the invention discloses a bearing evincing the features of c~aim 1. In a pertinent design of this bearing, optimal rcsults will be obtained with am active bearing defmed by the features of claim 11 and also in particular by means of the procedure defined in claim 12.
The sub-claims define further ~ "l,o.l - t~ of the invention.
In its simplest design, the bearing of the invention is a fluid bearing, more specifically an elastomer-amd-fluid bearing, wherein a support component resiliently rests on a matching rest component in particular in the form of a bearing housing by two physi-c lly separate elastomer springs which cooperate ~ ly amd in parallel, namely a radial and an axial spring. The terrninology "radial spring" amd "aAial sprmg" primarily relates to the position or orientation of their mam spring-force vector and only secondarily to thc geometric ""''-'L'""'"" and orientation of the bodies of the elastomer springs them-selves.
To elucidate such terminology, reference is made to the nume}ous conven-tional spring systems m can-shaped or cup-shaped bearing housirlgs of which the open end is sealed typically by a radial, matching, prestressed elastomer-spring disk. When this elastomer spring points ~radially" in relation to the bearing housing and the load, then this elastomer spring is crossed by an axial, mostly bar-shaped load comnector, usually being fastened to it by fusion, friction or geometric linkage. When such a bearing is statically or dy~ lly loaded in the axial direction, the radially pointing elastomer spring is compres-sively upset or tensively loaded in curving elongation. ln spite of their radial orientation, such elastomer-spring disks would be considered in the sense of tbis invention, because of the main vector of effective spring force, to be "axial springs"~ Such springs manifestly also act in centering maDner, but they evince hardly any useful radial spring-constant because they precisely achieve their maximum stiffness m the neutral plane wherein they are meant to be stress-free.
The bearing of the invention on the other hand shall be preferably designed in such marmer that the radial spring when im its neutral position be fully relaxed and ""~1. F~,, .., ~ This condition also can be provided for substantial static loads by using an matching prestressed axial spring which acts on the load connector just as the radial spring does.
In principle therefore t~vo elastomer springs or systems of elastomer springs are operative in the fluid bearing of the in~!ention, one of which provides axial ~ i .,. f.
and the other radial compliance for the dynarnic forces acting on the bearing. It is clear that for ~tance the radial spring will resiliently ~ ~eive not solely pure radial load ~

' ~ 2181619 but also to a slight extent axial ~ , and that the axial spring wil~ resiliently receive not only pure axial load ..i....l,...,. ,.l~, but also, even to a lesser degree, radial load compo-nents.
The ~ JlC~ Jlls "fluid", "fluid working chamber" or "fluid bearing" are deliberately so selected where they are used that the bearing design of the invention may be pneumatic or partly pneumatic or hydraulic or pardy hydraulic. The critical design of the invention is a fluid-damping bearing with two mutually separate but COu~ g elastomer springs or systems of elastomer springs, one of which is the radial spring and the other being an axial sprmg or equivalent ..,.1 IJ' ...1. -lly designed from the said radial spring and func-tionaUy separate from it. At least one wall of one of the t~vo springs .~ .,r.J..~ly operates as the e~astic wall of an operatingfluid chamber formed m the bearing.
In general this minimum of one working-fluid chamber is located around the axial spring and acts as an elastic working-chamber bounding waU with the radial spring.
In this bearing design and m one ~,.. 1~.1,.. 1 of the imvention, the radial sprmg comprises one or more sectors or boreholes running from the bearing-inside axialIy into the radial spring, said sectors or boreholes being sealed on the outside with Culll~ ,ly thin-walled elastomer .. ..l..~ These membr~me zones act as expansion 2~ones for the fl~ud, in particular for hydralllically decoupling dynamic loads of compara-tively high flcu~ ,;es but ~ow amplitudes. Such vibrations no longer are decoupled in the damping overflow duct to the balancing chamber. Foremost however this kind of expan-sion-zone decoupling achieves that, in particular regarding hydraulic /1.~collrlin~, the radial spring constant of the radial spring remains unaffected by the fluid, that is, no axial defor-mation of the radial spring takes place that otherwise would leaYe the fine control of the bearing in the radial direction to .,,.,.1.. ~ rather than to design.

'' ~ ?181619 In a ~ ,ul~ly significant design of the inverltion, the elastomer axial spring is a bello~vs spring. A "bellows" spring in this ~n~ denotes a hollow spring enclosed in at least a 5nhc~nti~11y cylindrical or more or less spherical marmer, and in particular it may also be a hy~ l)oloi~ or barrel-shaped, flat-walled or pleated or corrugated. The adYantage of such bellows springs, especially in the case of slightly convex cylinder walls, is fine control of bearing capacity with excellent radial strength due to its structure. In the pre~nce of large axial loads, good results are ~ lulG obtained in this case because the axial main load is absorbed by a steel helical screw mounted inside the elastomer bellows spring.
In such a case the bellows spring may evince lesser wall thickness and may be used for fine control and additional fluid decoupling and damping.
In orle ~.,.,I~r).l;..,..,l of the inverltion, the bellows spring is sealed ullll..d;l~,tiu~lly in fluid-tight manner and may be designed both as a self-balancing working chamber or with a ~parate balancing chamber connected to it. Preferably however the fluid system of the bellow spring will be a single chamber.
In an important design of the invention, the bellows spring acting as an axial spring or part of an axial-spring system m the form of a closed fluid chamber is fitted with a connection duct linking the fluid working chamber of the bellows spring to the arnbient.
The spring ~.1,,..,.. ~...~1.. ~ of the axial spring can be controlled m especially simple and rapid manner by means of such a ~ ,.I;rln duct, with a minimum of reaction delay, to oppose the dynamic load and hence to be ~ r., . ,l ,~"~,.l;, .~ Illustratively such activation of the bellows-spring fluid system can be ;...~ d by physically raising or lowering the hydraulic pressure of the bellows-spring working fluid, however it may also be carried out in particular by an electrically caused change in viscosity of the said working fluid and also s 218!6~9 irl especially sirnple amd rapid manner by illllll~ dial~ly arld directly applying acoustic waves to the working fluid.
The working-fluid pressure irl the bellows spring cam be controlled irl principle irl arbitrary manrler by means of the .-r)mm.mi.-~ti~.n duct, for instance usimg a corltrol valve, adjustabie pumps, membrane pumps or in the simplest wây also usir~g a piston.
Wherl usmg an electrically visco-elastic working fluid, the dampmg charac-teristics of the axial spring of the bearing also can be charlged relatively free of delays, in particular irl regulatirlg marmer, by applying electric fields while using throttling ducts or throttling apertures.
By means of said methods as well as by mearls of other cullc:~yolldillg methods, and usirlg a corltrol circuit of which the sensor tracks the amplitude of the dynamie load on t-he bearing, it is possible to stiffen the axial spring by applyirlg pressure, i.e. by loading it with a counter-force, in such a way that above-average dyrlamic loads applied to the bearing are balanced, so-to-speak "cancelled" by the introduced colmter forces.
The sub-c~aims contain further objects of the mvention.
The mvention is elucidated below by illustrative ,,.,I.o.l.",~ and in relation to the drawings.
Fig. 1 is a first embodiment of the invention partly shown in functional form -and in axial section, Fig~ 2 similarly shows a secorld ~",1,~.1.".. .,1 of the invention.
The bearing shown in Fig. I essentially consists of a load connector 1, a two-part bearing housing 2, 2" a~so acting by means of a matching rest pin 3 as a matching rest-fitting, of an elastomer radial spring 4 and of an elastomer axial spring 5.

~ 2 1 8 ! 6 1 9 The bearing comprises a frst fluid working chamber 6 and a second fluid working charnber 7.
The frst working chamber 6 is formed in a cylindrical bearing-housing slip-on ring 2" and is bounded at the load side by the elastomer radial spring 4 and on the match-mg-rest side by a matching rest bearing a plate 8 in turn clamped in fluid-tight manner between the two beafing housing parts 2, 2". In the f-.mho~:~fnt of Fig. 1, the elastomer radial sprmg 4 is integral with a continuous inside wall rubber coating of the bearing-housing slip-on ring 2'. The e~astomer is vulcanized in fused manner into the metal part.
Also the load connector I is fused into the elastomer radial spring 4.
An overflow duct 9 is forrned in the matching rest disk 8 and provides hydraulic f.l .. , .. ,:. ,.1 1.. , between the first working chamber 6 and a balancing chamber 10.
The balancing charnber 10 is bounded at the matching rest side by an easily deforming balancing membrane l l which is able to freely e~pand into the cup-shaped bearing-housing base 2.
The overflow duct 9 and the balancing chamber 10 are conventionally used to dampen large-amplitude load changes which are introduced at the load conmector I into the first hydraulic working chamber 6. Easily deforming thin-walled exparlsion zones 12 are used for hydraulic decoupling to darnpen ~u~ .,ly high-frefluency interfering vibrations of low amplitudes, said zones 12, depending on the design l~u,u..~ of the particular bearing ~rrlif~tifln~ may be formed in the elastomer radial spring 4. In the embodiment shown in Fig. 1, radial legs 2û are used to radially guide the suppûrted weight and the expansion ~;ones ~2 will be bridged by elastomer stops 21 in the case of radial impacts 218161q In the rl.ll,o.l;",~ .,t of the bearing of the invention shown in Fig. 1, the axial spring 5 is an elastomer bellows spring 13. Said spring 13 is inserted in ~ uc~l~;~lly locking manner into contoured recess on the base side of the connector I and on its end opposite the rnatchirlg-rest side is affixed in frictionally and ~...""~ lly locking manner to the matcning rest disk. The inside space of the bellows spring is thereby sealed in fluid-tight manner and serves as the second working chamber 7 for a second working-fluid system. The walls of the bellows spring 13 are c~u~ Liv~, y thin and easily d~
as a result of which a change in volume caused by dynarnically loading the bearing can be f~ 1 by a ~ ff~Tm~ti~-n of tne elastomer bellows spring 13 imto the first working chamber 6.
The second working chamber 7 may be designed both as a pneumatic spring and as a hydrf~n1ically damped spring. The expert will consider in malmer known per se the ~:yuil~,.l.~,llb of the field of application when selecting the axial-sprirlg rAnfi~lr~ifln In the r l,.ho~ .- ,1 shown in Fig. 1, the second, ie. the hydraulic fluid working charnber 7 is comnected to a duct 15 passing through the matching rest disk ~, tbrough the balancing chamber 10 and t-h-rough the balancing membrane 11 into the base 2 of the bearing housing. A supply conduit may be illustratiYely connected to the matching rest-side aperture of the duct 15 to aUow pressure control of the operational zone in the second working chamber 7 from an external fluid reservoir and an external pump. Similarly, electrical feeds rnay run through this duct which with 1,....".,~ . bearing design also may be used to electrically control the viscosity of a hydraulic working fluid.
In the ~mho~linn.~nt of Fig. 1, the matching rest-side aperture of the duct 15 is ,.1.. .~.,.li. ~lly covered by a pressure generator 16 sealing it in fluid-tight manner. The expression "pressure generator" denotes any known de~ice able to apply a pressure to a ~ 21 8 1 6 1 9 hydraulic working fluid, whether in the form of a shock wave or a vibration. For instance such a pressure generator may be a flexible control membrane, a ~ .. " .~ transducer, or merely a piston slidingly positioned in the duct segment ~ I;..g the cou~Jlil.g working cylinder.
In turn this pressure generator 16 is driven by an external control 17 connect-ed through a control line 18 to the pressure generator 16. The control 17 in turn is driven by serlsors 19 detecting loads applied to the bearing or, in the case of early detection, early pickup of the dynamic loads applied to the bearing As a result the control generates by means of the pressure generator a force component opposing said loads arld feeds said force in such manner into the working fluid in the second working chamber 7 that the dynamic loads acting on the bearing are balanced, that is are evenly cancelled.
Fig. 2 illustrates another ~ o-l;~ of the invention. Basically the bearing of Fig. 2 is that of Fig. 1, and therefore the references used in Fig. I also are used im Fig.
2.
Whereas in Fig. I the axial spring 5 comprises only one spring element, namely a bello~vs spring, that is, it consists of only one such, the axial spring 5 of the system shown in Fig. 2 is composed of two sprirlg elements, namely the bellows spring 13 of Fig. I and a steel-spring helical spring 14 mounted in said bellows spring. The elastomer bellows spring 13 is inserted in geometrically locking marlner m a base-side, contoured recess of the connector I in such manner that the bellows spring is made ~ ".. ,I, . y to the connector in order to form a geometrically locking snug-fit seat. At its opposite, matching rest-side end, the bellows sprirlg is affixed in frictionally and ~5~vl~ ally and also fluid-tight marlner to the matching rest disk 8 The spring-steel helical spring 14 is 2~81619 cl_mped inside tbe i .l,.. Ahly sealed bellows spring between tne load connector 1 and the matching-rest disk B.
As regards the GlllbOdilllCIl~ of the bearing of the invention shown in Fig. ~, the second working chamber 7 is designed in such a way that its operative fluid is the ambient air. Pressure balance is assured in this case through one or more _pertures A~2 in the bearing base 2 and the . 1,..".".,.~ n duct 15' leading to the ambient. In this design, the bearing shown in Fig. 2 is ~ lifi~lly more economical because of the absence of the electrorlic control and regulation of the inner fluid system of the bellows sprmg than the more elaborate and advanced system shown un Fig. 1.
Accordingly the bearing shown in Fig. 2 is ~ t ~,-1 as reg_rds damping in tbat it operates with a hydraulic fluid in the working chamber 6 and with a pneumatic working fluid, here the ambient air, in the irmer working chamber 7 in the bellows spring with.
Depending on the lc.lu..cl~ of the particular bearing arrli~ innc, also in the sense of matching the bearing data to the purpose of the ,.~ , the variations shown in Figs. I and 2 allow il~ il.g their particular elements. For inst~mce the bearing of Fig. I comprising the axial spring S in the form of the bellows spring may be additionally fitted with a helical spring or with an additional inner support rubber spring, especially a rubber spring buffer reinforced by shaped sheetmetal. Similarly the bearing shown in Fig. 2 comprising a 1~ a ~lly damping inner working chamber 7 in the axial spring system 5 may be designed to be solely a bellows spring without the support by a steei helical spring.
All the above ~..lbodi ~ and variations of Figs. I and 2 are ~,II~c~
by am especially modest required height and by marked .,--,.,1,~ Already in their u~ t~d passive design variations they allow wide-band decoupling between the load connector 1 and the matching-rest pin 3, such decoupling being offered by the Y~h~t~ntiAIIy larger pure hydraulic bearings of the state of the art on account of their more rigid, that is more ~ Li~lly conducting spring systerns.
The enclosed abstract of the invention is an integral part of the disclosure of the invention presented herein.
In this manner a bearing is created which darnpens over a wide frequency range as well in the range of impacts and which on account of the r~ A~ . of elastomer bellows spring and the steel-spring helical spring can be made unusually compact.

Claims (12)

1. A bearing, especially an engine bearing, to support in damping manner vibrating masses, composed of a load connector (1) engaged by an elastomer radial spring (4) and a separate elastomer axial spring (5) receiving the said load and braced for that purpose by a bearing housing (2, 2') serving as a matching rest, and further composed of at least one working chamber (6; 7) present in the bearing housing and fitted with a wall zone constituted by at least one of the two elastomer springs (4; 5).

2. Bearing defined in claim 1, characterized by a disk-shaped elastomer radial spring (4) sealing the bearing housing on the load-receiving side and into which is fused the load connector (1).

3. Bearing defined in claim 2, characterized by an easily deforming thin-walled expansion zone (12) in the radial spring (4).

4. Bearing defined in one of claims 1 through 3, characterized by an elastomer bellows axial spring (13).

5. Bearing defined in claim 3, characterized by an elastomer bellows spring (13) with a cylindrical or with a curved cylindrical axial outside wall.

6. Bearing defined in either of claims 4 and 5, characterized by an axial spring (5) in the form of an elastomer bellows spring (13) which is sealed on all sides in fluid-tight manner and where required merely evinces one or two apertures for filling, balancing, activation or venting a hydraulic or pneumatic working fluid.

7. Bearing defined in one of claims 1 through 6, characterized by a matching rest (8) supported in the bearing housing and by an axial spring (5) inserted in frictionally locking manner between the connector (1) and the matching rest (8) inside an enclosing working chamber (6) holding a damping fluid.

8. Bearing defined in claim 7, characterized by a matching rest (8) of which the load-side surface forms one of the boundary walls of the working chamber (6) and of which the opposite, rest-side surface forms one of the boundary walls as a wall segment of an easily deforming membrane (11) containing a damping fluid balancing chamber (10), a damping overflow duct (9) being present in said balancing chamber (10) through which both chambers (6, 11) communicate with each other.

9. Bearing defined in one of claims 4 through 8, characterized by a steel helical spring (14) inserted in frictionally locking manner into the elastomer bellows spring (13) between the load connector (1) and the matching rest (2).

10. Bearing defined in one of claims 4 through 9, characterized by a connection duct (15) present inside the bellows spring (13) and communicating with the ambient and by means of which the bellows spring can be loaded through its fluid working medium with impact-damping and vibration-damping counter-forces in the form of impacts or pressure vibrations.

11. Bearing defined in claim 10, characterized by a pressure generator (16) coupled outside the bellows spring (13) to the operating-fluid connection channel (15) pressurizing said working fluid, said generator in turn being driven by a control member (17) in relation to the instantaneous operational load on the bearing.

12. A method for damping and extinguishing impacts and vibrations of a dynamically loaded pneumatic bearing or a hydrodynamically loaded bearing, characterized in that in relation to the instantaneous operational parameters of the dynamically loaded bearing, forces generally in the form of impacts or vibrations are impressed on the working fluid of the bearing in at least one of the possibly several working chambers in order to counter the instantaneous load on the bearing.