GB2285851A - Bridging clutch friction ring - Google Patents

Abstract

A friction ring for use in a wet-type clutch, more particularly for a bridging clutch 15, Fig. 1, of a hydrodynamic torque converter 3 has a friction face 21 with an outer circumference and an inner circumference. Grooves such as 24, 26, Fig. 2, are provided in the surface of the friction face 21 for cooling. These grooves ensure a connection between the outer circumference and inner circumference. The grooves form over part of their extended length at least one throttle point 30 which determines the volume of cooling fluid which can flow through the grooves. <IMAGE>

Description

Friction ring and clutch with said friction ring The invention relates to a friction ring or friction lining for use in conjunction with a wet-type clutch, more particularly for use in a bridging clutch of a hydrodynamic torque converter wherein the friction ring forms a friction surface which is provided with grooves or channels for carrying a cooling fluid.

Friction rings or friction linings of this kind as well as wet-type clutches fitted with same are known through US-PS 4 969 543 and 5 056 631. These two US patents describe hydrodynamic torque converters with a bridging clutch wherein the engaging friction faces are designed so that even when the bridging clutch is closed an air flow is possible between the chambers provided each side of a ring piston. These hydrodynamic torque converters have a housing containing a pump wheel, turbine wheel, guide wheel and bridging clutch which has the ring piston. Either side of the ring piston are formed the oil-fillable chambers wherein the first of the chambers is formed radially inside the friction faces of a bridging clutch and at least the turbine wheel is provided in the second chamber. The first chamber is defined by the ring piston and a radial wall of the housing. The oil stream serves to reduce the thermal strain on the components which occurs as a result of slip in the bridging clutch, more particularly in the area of the friction lining or friction faces.

The object of the present invention is to improve the known friction rings having grooves for conveying a cooling fluid as well as to improve the wet-type clutches equipped with same with regards to the cooling effect which can be produced by the volume flow of fluid. This is to be achieved inter alia through an improved heat exchange in the area of the friction faces of the wet-type clutch between the fluid and the adjoining components. Furthermore a high torque capacity of the wet-type clutch is to be achieved through the design of the friction ring or wet-type clutch according to the invention. Furthermore the friction ring or the friction disc fitted with said friction ring and thus also the wet-type clutch should be capable of simple and economic manufacture.

According to the invention this is achieved in that in the area of the friction face of the friction ring there are cooling grooves which are designed so that they form at least over part of their length extension at least one throttle point which mainly determines the volume of cooling fluid which can flow through the grooves. The grooves can thereby be designed so that they guarantee a connection between the outer diameter or the outer circumference and the inner diameter or inner circumference of the friction ring. The grooves are thus open radially towards the outside and radially towards the inside. It is particularly advantageous if the at least one throttle point is designed so that a turbulent flow is produced in the area of same.

The groove areas which are located outside of such a throttle point should be designed with regard to the width and depth so that an at least substantially laminar flow occurs there.

Through the design and arrangement of the cooling grooves or cooling channels according to the invention it is possible to achieve a direct cooling of the engaging friction faces, more particularly in the case of permanent slip. Through the throttling of the volume flow of cooling fluid according to the invention in dependence on the differential pressure between the two chambers of a converter bridging clutch either side of a piston it is possible to achieve an optimum cooling over the entire operating area of the corresponding hydrodynamic torque converter. The design according to the invention has the advantage that in the area of the throttle points there is a comparatively high pressure drop as a result of the turbulent flow which exists there whilst in the remaining groove areas the flow losses are very slight as a result of the comparatively large through-flow crosssection which is present. With the prior art already mentioned the grooves are designed so that mainly a laminar throttling exists here over the entire length. With this type of throttling the volume flow rises linearly with the pressure or linearly with the pressure difference between the two chambers of the bridging clutch. With a turbulent throttling according to the present invention the volume flow rises according to a root function in dependence on the existing pressure or the existing pressure difference. This means that a turbulent throttling is more favourable since practically over the entire pressure range which occurs with a hydrodynamic torque converter a larger volume flow is provided below the maximum permissible pressure.

By using according to the invention throttle points with a comparatively small length in relation to the overall length of a groove it is also possible to keep to a low level or to minimize the effect of the manufacturing tolerances of the lining grooves (width and depth) as well as the effect of the manufacturing tolerances and operationally-conditioned distortions of the piston and counter friction faces on the flow resistance of the lining grooves. Since the oil viscosity decreases as the temperature rises it is also possible through the throttle points according to the invention to achieve a greater volume flow and thus improved cooling as the temperature of the cooling oil increases The flow resistance of the throttle points or lining grooves thus decreases in relation to the oil with increasing temperature. This effect which is positive in itself must however be limited to the desired amount of oil volume through a corresponding design of the throttle points since with too much through flow the pressure in the closing chamber of the bridging clutch can no longer be maintained.

It can be particularly advantageous if the throttling of the cooling means takes place in diaphragms, that is short channel parts with sharp-edged flow inlet and/or flow outlet. A satisfactory turbulent flow can thereby be guaranteed whereby the flow resistance is each time only dependent linearly on the width and depth of the diaphragm or throttle point. With long channels with a substantially laminar flow, as is the case for example in-the prior art listed, the flow resistance is dependent in the fourth power on the hydraulic radius or diameter. This means that tolerances of the groove dimensions very severely affect the flow resistance.

The grooves or channels can be formed in the friction face of the friction ring or friction lining so that the friction face forms radially outwards a practically continuous ring area which is only broken by the throttle channels which run radially or inclined. Radially inside the ring area there are groove areas or channel areas with a considerably larger cross-section compared to the cross-section of the throttle channel so that in these areas there is an only very much lower flow resistance compared to the flow resistance provided in the throttle channels. The main part of the overall flow resistance of the cooling grooves is thus present in the area of the throttle points or throttle channels.

The throttle points are preferably mounted on the outer diameter of the bridging clutch or friction ring since the contact pressure force or closing force of the bridging clutch can be increased thereby. This is due to the fact that over the largest part of the radial extension of the engaging friction faces there is a pressure throttled to a substantially lower pressure level whereby the closing force can be correspondingly increased.

It is particularly advantageous if the throttle points are arranged so that they always lie in a supporting area of the friction face so that the throttling cannot be by-passed through a gap between the lining surface and the surface of the counter friction face. For this reference is made to Figures 7 and 8. Provided the remaining lining areas in which the channel sections have a comparatively large crosssection do not fully adjoin the counter surface and are thus flooded the overall flow resistance is reduced to an amount which does not impair the function since the channel sect ions provided here only undertake a very small part of the overall throttling. The cooling oil grooves or channels are preferably made comparatively deep whereby the effect of the manufacturing tolerances and setting of the lining on the flow resistance is minimized. The channel guide is to be carried out so that no dead areas exist which cannot be passed. The grooves provided in the friction face or in the friction lining can be indented or punched through.

The length of a throttle point can advantageously be between 2 and 8 mm, preferably in the order of between 3 and 5 mm.

The cross-sectional ratio between the oblong areas of the grooves with larger cross-section and a throttlecpoint can be in the order of between 3 to 1 and 8 to 1, preferably in the order of between 4 to 1 and 6 to 1. However larger or smaller ratios are possible depending on the type of use.

Further expedient embodiments of the invention will be apparent from the sub-claims and the following drawings.

The invention will now be explained, by way of example, with reference to Figures 1 to 11 in which: Figure 1 is a sectional view through a device with a wet-type clutch having a friction ring according to the invention; Figure 2 is a partial view of a friction lining designed according to the invention; Figure 3 is a sectional view on an enlarged scale according to line III of Figure 2; Figure 4 is a sectional view along the line IV of Figure 2; Figure 5 shows the pressure distribution in the radial direction which can be achieved through the arrangement according to the invention of a throttle point according to the invention in the area of the friction face or in the grooves; Figure 6 is a diagram for explaining the effect of the improved groove design; Figures 7 and 8 show different arrangements for a friction lining and Figures 9 to 11 show further design possibilities for cooling channels or cooling grooves.

The device illustrated in Figure 1 has a housing 2 containing the hydrodynamic torque converter 3. The housing 2 is connectable with a driving shaft which can be formed by the output shaft of an internal combustion engine such as eg the crankshaft. The rotationally fixed connection between the driving shaft and the housing 2 can be provided by a drive plate which is connectable rotationally secured radially inwards with the driving shaft and radially outwards with the housing 2. A drive plate of this kind is known for example from JP-POS 58 30532.

The housing 2 is formed by a housing shell 4 adjoining the driving shaft or internal combustion engine, and by a further housing shell 5 fixed on same. The two housing shells 4 and 5 are joined together with a fixed sealing connection radially outwards by a welded connection 6. In the illustrated embodiment the housing shell 5 is used directly to form the outer shell of the pump wheel 7. To this end the blade plates 8 are fixed in known way on the housing shell 5. The housing shell 5 is pushed axially onto the outer sleeve-like area 4a of the housing shell 4. A turbine wheel 10 is provided axially between the pump wheel 7 and the radial wall 9 of the housing 4 and is connected fixed or rotationally rigid with an output hub 11 which can be coupled rotationally secured with a gearbox input shaft through internal gearing. A guide wheel 12 is provided axially between the radially inner areas of the pump and turbine wheels. The housing shell 5 has radially inwards a sleeve-like hub 13 which can be mounted rotatable and sealingly on the housing of a gearbox. In the inner space 14 formed by the two housing shells 4,5 is a bridging clutch 15 which is mounted actively parallel with the torque converter 3. The bridging clutch 15 allows a torque coupling between the output hub 11 and the driving housing shell 4. A rotationally elastic damper 16 is connected in active engagement in series with the bridging clutch 15 and in the illustrated embodiment is provided between the ringshaped piston 17 of the bridging clutch 15 and the output hub 11. The rotationally elastic damper 16 comprises in known way energy accumulators in the form of coil springs.

The ring-like piston 17 mounted axially between the radially aligned wall 9 and the turbine wheel 10 is mounted radially inwards on the output hub 11 where it is displaceable axially to a limited extent. The ring-like piston 17 divides the inner space 14 into a first chamber 18 which is formed radially inside the friction engagement area 19 of the bridging clutch 15 axially between the ring like piston 17 and the radially housing wall 9, and a second chamber 20 in which inter alia the pump wheel 7, turbine wheel 10 and guide wheel 12 are all housed.

The housing shell 4 forms with a ring-like radially outer area a friction face 21 which can be brought into friction engagement with a friction lining 22 which is supported by the ring-like area 23 of the piston 17.

In more recent concepts for a drive train, eg of a motor vehicle the bridging clutch is operated with slip over at least a large part of the operating area of the torque converter whereby during the slip phases in the friction engagement area 19 a loss of power in the form of heat occurs which in certain operating conditions can be very high and amount to several kilowatts. Such operating states exist for example when climbing uphill with trailers or when changing from a non-bridged state to a practically bridged state of the converter clutch. Such concepts for operating a converter bridging clutch with slip have been proposed for example by German Patent Application P 43 28 182.6.

In order to avoid unacceptably high temperatures in the friction engagement area 19, and thus to counteract a destruction of at least the friction lining surface and a part of the oil present in the inner space 14, in the illustrated embodiment means are provided in the form of oil grooves or channels 24 set in the friction lining 22, through which even when the bridging clutch 15 is practically closed a constant oil flow can be produced between the second chamber 20 and the first chamber 18. The oil flow is thereby directed over the friction surface 22a of the friction lining 22 and the friction surface 21. The oil channels 24 are improved with regard to their shape so that a good heat exchange can take place between the components producing the friction engagement in the area 19, and the oil flowing through. A preferred shaping of the channels 24 is described in further detail in connection with Figures 2 to 4.

The end of the channels 24 lying radially further outwards is connected with the chamber 20 and the end of the channels 24 lying radially further inwards is connected with the chamber 18. When the bridging clutch 15 is closed the cooling oil stream flows over the channels 24 into the chamber 18 and here radially towards the rotary axis 25.

This cooling oil stream can then be diverted off in the area of the output hub 11, for example through a hollow shaft or through a channel provided for this purpose, namely preferably at first into an oil cooler. From this oil cooler the oil can be returned to a sump and from there back again into the hydraulic regulating or control circuit.

Figure 2 shows part of a circular ring-shaped friction lining 22 which can be used in the case of a converter bridging clutch according to Figure 1. The friction lining 22 has circumferentially distributed grooves or indentations 26 which form the connecting channels 24 between the two chambers 18 and 20.

The friction lining 22 has an outer circumference or outer diameter 27 as well as an inner circumference or inner diameter 28. A short partial section 29 of a channel 24 connecting the outer diameter 27 with the inner diameter 28 forms a throttle point or throttle diaphragm 30. The partial section 29 of a channel 24 is thereby radially aligned and changes radially inwards into circumferentially aligned radially outer channel sections 31 which pass through hairpin type deflections 32 into radially further inner channel sections 33 which are likewise circumferentially aligned. The channel sections 33 are connected with a radially inwardly open outlet area 34 for the cooling means which is directed over the channels 26.

The channel areas or channel sections 31, 32, 33 and 34 adjoining a throttle point 30 are designed in cross-section in relation to the cross-section of a throttle point 30 so that here there is practically mainly a laminar flow even with the maximum differential pressure occurring between the two chambers 18 and 20 of the device according to Figure 1.

A turbulent flow practically always takes place in the area of a throttle point 30 during operation of the device according to Figure 1 and in the event of friction engagement of the bridging clutch 15. The channels 24 are thus designed so that the volume of cooling fluid flowing through same is not determined by the flow resistance produced over the entire length of the channels, as is the case with the known prior art, but is mainly determined by the resistance existing in the area of the throttle point or throttle points 30. As can be seen from Figure 2 the channel sections 31 and 33 running in the circumferential direction have partial sections 35, 36 which in relation to a throttle point 30 extend in different rotary directions.

The partial sections 35, 36 are thereby mounted symmetrical relative to a throttle point 30 - viewed in the circumferential direction.

As can be seen from Figures 2 to 4 the expansion 29 of a throttle point 30 only amounts to a very slight part of the overall length of a channel 24. For the usual structural sizes of wet-type clutches or bridging clutches 15 with an outer friction diameter 27 in the order of between 180 and 260 mm the length 29 of a throttle point 30 can amount to between 2 and 8 mm, preferably between 3 and 5 mm, depending on the type of use.

The through-flow cross-section of a throttle point 30 is smaller by a multiple in relation to the through-flow crosssection of the groove sections 31, 32, 33, 34 adjoining the outlet side of the corresponding throttle point 30. The ratio can thereby be in the order of between 1 to 2 to 1 to 10. For most cases however a ratio is sufficient of between 1 to 4 and 1 to 6.

As a result of the substantially greater through-flow crosssection of the channel sections 31, 32, 33, 34 it is ensured that in these sections a laminar flow takes place practically always or in the main.

In order to achieve an optimum turbulent flow in the area of the throttle points 30 formed by short channel-type indentations it is expedient if at least the cross-section in the inlet area of the throttle points 30 is sharp-edged.

In the illustrated embodiment according to Figure 2 the throttle points 30 change on the outlet side through a gradual expansion formed by rounded areas into the corresponding groove areas 31. It can however be expedient if a sharp-edged cross-sectional change is provided between the channel sections 31 and the throttle points 30.

The surface part of the surface claimed by the grooves or channels 24 can be in the order of between 30 and 65 percent, preferably in the order of between 40 and 55 percent in relation to the surface existing between the outer diameter 27 and inner diameter 28. In the illustrated embodiment according to Figure 2 this proportion amounts to about 50 percent.

The throttle points 30 are preferably designed so that through these about 60 to 85 percent, preferably 70 to 80 percent of at least the maximum pressure difference between the two chambers 18 and 20 is broken down. This means that after the throttle points 30 or shortly behind the throttle points 30 the pressure existing in the channel sections 31 is only still larger by about 15 to 40 percent or 20 to 30 percent than the pressure in the chamber 18. As a result of the action of the throttle points 30 it is expedient if as shown in Figure 2 these throttle points are mounted in the outer area or on the outer radius of the friction ring 22, thus in the area of the higher pressure since the pressure building up in the area of the friction faces 21, 22a and counteracting the closing pressure of the bridging clutch 15 can be kept small. The torque transferrable by the bridging clutch 15 for a pressure difference between the two chambers 18 and 20 can thereby be increased compared to the previously known bridging clutches with cooling channels and a corresponding volume of cooling fluid. By shifting at least some of the throttle points 30 radially inwards it is possible however to reduce also the torque capacity of the bridging clutch 15 for a given differential pressure between the two chambers 18 and 20.

From Figure 5 can be seen the effect of the radial arrangement of the throttle points 30 on the transferrable torque. In Figure 5 a partial area of the housing shell 4 and of the piston 17 with friction lining 22 fixed thereon is shown on an enlarged scale on the left hand side.

Possible idealized pressure profiles over the radial extension area of the friction lining 22 and in dependence on the arrangement of the throttle points are shown on the right hand side of Figure 5. For a given higher pressure pl in the chamber 20 and a given lower pressure p2 in the chamber 18, viewed over the radial extension of the friction lining 22, with an arrangement of the throttle points 30 radially outwards, as is the case in Figure 2, in the area between the friction face 22a of the friction lining 22 and the friction face 21 a pressure distribution is possible in the channels 24 running according to the chain-dotted line 37. From the chain-dotted line 37 it can be seen that in the area of the throttle points 30 about 80 percent of the pressure difference between pl and p2 is dissipated. The difference between the pressure Pa which exists close to the outlet side of the throttle points 30 and the pressure p2 in the chamber 18 is thus comparatively small. With the arrangement of the throttle points 30 radially inwards, thus in the area of the outlet sections 34 according to Figure 2, a pressure distribution would arise in the engagement area 19 according to the dashed line 38. From the two lines 37 and 38 it can be seen that for a given pressure difference between the two chambers 18 and 20 the torque transferrable by the bridging clutch 15 can be affected by the arrangement of the throttle points 30 on different diameters. By arranging the throttle points 30 radially outwards the differential pressure between the two chambers 18 and 20 required to transfer a certain torque can be reduced relative to the previous bridging clutches with a cooling oil stream between the two chambers 18, 20. Depending on the number and shape of the throttle points 30 the throughflow width of such a throttle point 30 can be in the order of between 0.4 and 2. 5 mm, preferably in the order of between 0.5 and 1.5 mm. The depth of the grooves 26 can be in the order of between 0.2 and 1 mm, preferably in the order of between 0.3 and 0.7 mm. The depth of the grooves 26 can be practically the same over the entire extension thereof. The grooves 26 can however also have areas of different depth. More particularly in the area of the throttle points 30 as well as where applicable in the transition area between a throttle point 30 and the remaining groove sections 31 it can be advantageous if a greater depth is present. This is shown in Figure 3 by the chain-dotted line provided with reference numeral 30a. It can thus be advantageous if in order to keep the desired through-flow cross-section of a throttle point 30 the throttle point is set lower compared to the other groove areas and is designed smaller in width to compensate. It can thereby be guaranteed that the dependence of the throttle action of a throttle point 30 in relation to the wear on the friction lining 22 which causes a crosssectional reduction of the throttle point 30 is reduced.

According to the invention the volume flow of cooling fluid inside the wet-type clutch is thus set by means of at least one throttle 30 wherein comparatively long channels can be formed in the remaining lining surface - viewed in the flow direction - behind the corresponding throttle point 30, thereby guaranteeing the smallest possible flow resistance and a large heat-exchanging surface.

The pressure difference pl - p2 (ap) between the two chambers 20 and 18 is shown in Figure 6 on the abscissa axis. The volume flow which is set in dependence on the pressure difference available is shown on the ordinate axis.

With a laminar throttling of the volume flow over the length of the grooves incorporated in a friction lining there is a practically linear connection between the pressure difference at the grooves and the volume flow. This connection is represented by the straight solid line of Figure 6. The difference between the pressure on the input side and the pressure on the output side of the corresponding groove or grooves is taken as the pressure difference at the grooves. A laminar throttling of this kind is produced in practice when the grooves are designed according to the prior art mentioned at the beginning, namely US-PS 4 969 543 and 5 056 631. With this prior art the proportion of laminar throttling amounts to about 70 percent of the overall throttling taking place in the channels.

The dotted line represents the volume flow which can be achieved through turbulent throttling according to the invention. The volume flow path in dependence on the pressure difference between the two chambers 20, 18 corresponds substantially to the path of a root function.

The dotted path can be achieved through the design of the grooves according to the invention, more particularly according to Figure 2. As can be seen from Figure 5, particularly with smaller pressure differences with a turbulent throttling there is a greater volume flow than with a laminar throttling. This is particularly advantageous since with the lock-up even with small pressure differences between the two chambers 20 and 18 there should be the largest possible volume flow available in order to ensure the best possible cooling action.

The groove designs corresponding to the two characteristic lines according to Figure 6 are designed so that they ensure the same volume flow for a predetermined Ap max. This Ap max lies with the usual hydrodynamic torque converters with bridging clutch in the order of between 7 and 10 bar. The Ap max can however also be below or above this band width.

The design according to the invention of the grooves or channels provided for a cooling oil flow furthermore makes it possible to reduce the temperature dependence of the through-flow through these channels, namely because the main throttling, thus the main pressure drop takes place in the area of the comparatively short throttle points. The groove in the area of a throttle point enters only linearly into the through flow or volume flow whereby a lower dependence on the geometric tolerances is ensured. With the design of the grooves according to the said prior art the main part of the throttle takes place laminarly namely over the entire length of the grooves. With such a throttling the groove height enters in the fourth power into the through flow or into the volume flow. This results in a strong dependence regarding the geometric tolerances of the lining of grooves.

Furthermore as a result of the laminar throttling present there is a strong dependence of the volume flow on the viscosity or temperature of the cooling medium.

In order to ensure that the grooves according to the invention always ensure their throttling function it is necessary for the friction lining 22 to adjoin the counter friction face 21 in the area of the throttle points. It should at least be guaranteed that no gap opens up in any of the operating states arising in the area of the throttle point or if so such a gap should not be greater than 0.03 mm, preferably no greater than 0.01 mm. Such gaps can arise as a result of insufficient parallel alignment between the friction faces which can be brought into engagement.

In order to ensure that in all operating states in which the friction faces engage the throttle points 30 carry out their function it is advantageous if the friction lining 22 is supported according to Figure 7 by a component, namely the ring piston 17.

A partial area of the housing shell 4 as well as of the piston 17 with the friction lining 22 fixed thereon is shown on an enlarged scale in Figures 7 and 8. Figure 7 shows the shape of the piston 17 which said piston takes in the practically unstressed thus relaxed state. This piston shape is produced when there is practically the same pressure or only a comparatively slight pressure difference in the two chambers 18 and 20. In the relaxed state of the piston 17 the outer area 17a which holds the friction lining 22 is designed so that the friction face 22a of the friction lining 22 and the friction face 21 of the housing 4 include between them a wedged-shaped air gap 39 which expands radially inwards and has an angle which can be in the order of 0.5 and 3 degs., preferably in the order of 1 deg.

Figure 8 shows the position of the piston 17 which this occupies when there is a pr in the chamber 20 increases relative to the chamber 18 the piston 17 is deformed from the shape shown in Figure 7 into the shape shown in Figure 8. The contact area between the friction faces 21 and 22a thereby gradually increases or the angle existing between the friction faces 21 and 22a becomes smaller. The throttle points 30 also guarantee however a satisfactory control of the volume of cooling fluid.

The friction lining shown in Figure 2 or the friction ring 22 is formed in one piece. This could however also be comprised of several circumferentially adjoining sectorshaped individual lining parts.

Figures 9 and 10 show in part the friction linings 122, 222, 322 which are fitted with grooves or with channels according to the present invention It is common to the friction linings according to Figures 9 to 11 that they have throttle points 130, 230, 330 which are spread out over the circumference of the friction lining.

These throttle points 130, 230,330 mainly determine the volume flow which can pass through the channels 124, 224, 324. The channel sections adjoining the throttle points 130, 230, 330 have a substantially larger through-flow cross-section than the throttle points 130, 230, 330 so that there is mainly a laminar flow in these channel sections.

The flow speed in these partial sections of the channels 124, 224, 324 is thereby considerably less than the flow speed in the throttle points 130, 230, 330. This produces the optimum heat transfer between the cooling medium or cooling oil flowing through and the adjoining component parts.

With the design according to Figure 9 the friction lining 122 has a ring-like groove 131 which is connected with the throttle points 130 and which in turn is connected with a number of inwardly running radial grooves 132. The friction face of the friction lining 122 is formed by the protrusions 132a provided between the individual grooves 132, and the ring-like protrusion 122a which is provided on the edge area of the friction ring 122 and which is divided by the throttle points 130 into individual sector-shaped sections.

The friction lining 222 according to Figure 10 has a number of ring-like indentations 231, 231a, 23lb which are connected together by radially aligned groove areas 232, 232a. The radially inner ring-like groove area 231b is opened radially towards the inside through radial groove areas 232. The radial groove areas 232, 232a and 232b are off-set circumferentially relative to each other so that a multiple deflection of the oil flowing through the channels 224 takes place.

With the embodiment illustrated in Figure 11 the channels 324 adjoining the throttle points 330 are formed circumferentially meander-shaped so that as a result of the surface and length of the meander-shaped areas of the channels 324 there is a good heat exchange between the cooling oil and the adjoining components or the adjoining friction faces.

According to a further design possibility of the invention the cooling grooves designed according to the invention can be provided in the area of the friction face 21 of the housing 4 instead of being formed in the friction lining 22.

These cooling grooves can be formed by imprinting in the sheet metal material. The imprinted channels have to be designed radially outwards and radially inwards so that they are open towards the chamber 18 and 20. Furthermore, instead of being supported by the piston 17 the friction lining 22 can also be fixed on the housing 4. Furthermore a friction lining 22 can be supported by an intermediate lamellar plate as is the case for example in some embodiments in the said prior art. The cooling channels according to the invention can furthermore be imprinted directly into the material forming the piston 17 wherein then the friction lining 22 is supported by the housing 4 or by a intermediate lamellar plate.

The grooves or channels formed in a friction lining or friction ring can be produced during manufacture of the friction lining, thus before fixing the friction lining on the support component, such as for example a ring piston or lamellar plate. The grooves, indents or channels can however also be formed in the friction lining during fixing, for example by sticking the friction lining on a support component, or after such fixing. Thus the friction lining, for example 22 according to Figure 2, can first be fixed on the ring piston 17 and during or after this fixing the channels 24 can be imprinted in the friction ring 22. The latter is carried out by means of a pressing tool which has suitable profiled areas.

The invention is not restricted to the embodiments described and illustrated but also includes variations which can be formed in particular by a combination of individual features, elements and functions described in connection with the different embodiments.

The applicant retains the right to claim as being essential to the invention further features only disclosed up until now in the description.

Claims (20)

PATENT CLAIMS

1. Friction ring for use in a wet-type clutch, more particularly for a bridging clutch of a hydrodynamic torque converter wherein the friction ring has at least one friction face with an outer circumference and an inner circumference, grooves being provided in the area of the friction face for cooling, these grooves ensuring a connection between the outer circumference and inner circumference, characterised in that the grooves form over part of their extended length at least one throttle point which determines the volume of cooling fluid which can flow through the grooves.

2. Friction ring according to claim 1, characterised in that the throttle point is designed for a turbulent flow and the remaining groove areas are designed for a substantially laminar flow.

3. Friction ring according to claim 1 or 2 characterised in that the length of one throttle point is in the order of between 2 and 8 mm, preferably in the order of between 3 and 5 mm.

4. Friction ring according to one of claims 1 to 3 characterised in that the cross-sectional ratio between the oblong areas of the grooves with larger cross-section and the throttle point lies in the order of between 3 to 1 and 8 to 1, preferably in the order of between 4 to 1 and 6 to 1.

5. Friction ring according to one of claims 1 to 4 characterised in that the throttle point is formed by a short channel-type indentation with sharp-edged flow inlet and/or flow outlet.

6. Friction ring according to one of claims 1 to 5 characterised in that the at least one throttle point is provided on the outer circumference of the friction ring.

7. Friction ring according to one of claims 1 to 6 characterised in that the friction ring has a number of radially aligned throttle points spread out over the circumference starting from the outer circumference and passing into circumferentially aligned groove sections which are connected radially inwards with a drainage section open towards the inner edge of the friction ring.

8. Friction ring according to claim 7 characterised in that the throttle points change into a radially outer circumferentially aligned groove section which is connected by radially aligned groove sections with an inner circumferentially aligned groove section which opens into a drainage section.

9. Friction ring according to claim 7 or 8 characterised in that the circumferentially aligned groove sections are arranged symmetrically relative to the associated throttle point - seen in the circumferential direction.

10. Friction ring according to one of claims 7 to 9 characterised in that - viewed in the radial direction - a drainage cross-section lies opposite a throttle point.

11. Friction ring according to one of claims 1 to 6 characterised in that a number of throttle point s are provided spread out over the circumference of the friction ring passing into groove sections which are guided zig-zag fashion or meander fashion in the circumferential direction.

12. Friction ring according to at least one of the preceding claims, characterised in that the grooves have at least two deflections.

13. Friction ring according to one of claims 1 to 12 characterised in that in relation to the surface provided between the outer circumference and inner circumference of the friction ring the surface portion which is included by the grooves lies in the order of 30 to 60 percent, preferably in the order of 40 to 50 percent.

14. Friction ring which is a component part of a bridging clutch of a hydrodynamic torque converter wherein the torque converter has a housing containing a pump wheel, turbine wheel, guide wheel and bridging clutch, the bridging clutch has a ring piston either side of which is an oil-fillable chamber, the ring piston supports at least one friction surface which can be brought into friction engagement with a counter friction face wherein the first of the chambers is formed radially inside the friction faces between the ring piston and a component part supporting the counter friction surface and at least one of the friction faces is formed by a friction ring according to claims 1 to 13 wherein as a result of the pressure difference existing between the two chambers an oil flow can be produced through the grooves formed in the friction ring in the event of axial contact of the friction faces.

15. Friction ring according to claim 14 characterised in that in the area of a throttle point the pressure difference which exists between the two chambers is reduced by about 60 to 90 percent, preferably 70 to 80 percent.

16. Friction ring according to one of claims 14 to 15 characterised in that the throttle points adjoin the chamber which when the bridging clutch is closed has the higher pressure.

17. Friction ring according to one of claims 1 to 16 characterised in that the grooves are formed by indentations or cut-out sections in the friction ring.

18. A wet-type clutch, more particularly for a bridging clutch of a hydrodynamic torque converter including a friction ring which has at least one friction face with an outer circumference and an inner circumference, wherein grooves are provided in the area of the friction face for cooling, these grooves ensuring a connection between the outer circumference and inner circumference, and the grooves forming over part of their extended length at least one throttle point which determines the volume of cooling fluid which can flow through the grooves.

19. Friction ring for use in a wet-type clutch substantially as herein described with reference to the accompanying drawings.

20. A clutch substantially as herein described with reference to the accompanying drawings.