GB2316153A - Friction lining having grooves through which cooling oil is directed - Google Patents

Abstract

A friction lining 422, for a clutch of a torque converter, has at least one circuitous grooves 435 which directs hydraulic fluid from a high pressure cavity across a width of the friction lining, in at least three traversing passes, to a low pressure cavity. The grooves have an inlet section 439 that opens at an outer edge of the friction lining to the high pressure cavity, and an outlet section 440 that opens at an inner edge of the lining to a low pressure cavity, the outlet section 440 is angularly offset relative to the inlet 439. Grooves 435 of this form are of a long length which ensure good heat exchange and thereby a low thermal strain on the linings 422. The grooves may be punched-out of the linings 422 material. Friction lining 422 has a radially outer cohesive area 422a and a cohesive radially inner area 422b each extending in a circumferential direction.

Description

Friction Clutch for a Torque Ccnverter The invent on relates to a friction clutch for a hydrodynami c torque converter which has a pump wheel, a turbine wheel, a guide disc and a lock-up friction clutch.

The clutch has an annular piston contained in a housing, and either side of the piston there is a chamber which can be filled with oil. The piston supports at least one friction face which an be brought into friction engagement with a counter friction face wherein the first cf the chambers is formed radially inside the friction faces between the annular piston and a component part supporting a counter friction surface, and furthermore channels or openings or ports are provided in at least one of the components forming or supporting the friction faces to produce an oil flow from the second of the chambers through the channels or openings or ports radially inwards towards the axis of rotation of the torque converter.

Through EP 0 078 651 a torque converter is known with a lock-up clutch wherein on the side of the ring piston remote from the frIction lining or friction face are channels which are connected through openings on one side with the first chamber formed axially between a radial wall of the housing and the rig piston and on the other side with the second chamber housing the turbine and pump wheel. An oil stream is providec through the channels from tne second chamber into the first chamber and serves to cool the viscous clutch which is provided in the torque flow between the ring piston and turbine hub.

Through US-?S 4 969 453 hydrodynamic torcue converters with lock-up clutches are known wherein the ring piston in the radial area of its friction face or the friction lining interacting with same has channels which even when the lockup clutch is closed allow an oil flow from the second chamber housing at least the turbine wheel into the first chamber which is defined by the ring piston and a radial wall of the housing. The oil flow thereby semes to reduce the thermal strain on the components, more particularly in the area of the friction lining or friction faces, which occurs as a result of slip in the lock-up clutch.

A further lock-up clutch for a hydrodynamic tcrque converter with channels provided in the friction face or in the friction lining is known from JP-OS 58-30532.

Furthermore it is known to operate converter lock-up clutches with slip wherein this slip can appear temporarily, eg in the case of gear changes, or practically over the entire operating range of the torque converter (permanent slip), depending on the design of the drive train and/or depending on the gear phase engaged and/or on the operating state of the drive interacting with the torque converter.

During the slip phases a loss of power in the form of heat occurs in the area of the friction lining or friction faces and under certain operating conditions this heat loss can be very high and amount to several kilowatts. Such operating conditions are for example when climbing 'pill with a trailer wherein a high power loss can take place over a longer time, and when changing from the non-bridged to the bridged state o the converter clutch, whereIn as a result of the temporarily high slip a very high power loss or heat loss can occur within a short time span.

As was explained with reference to the prior art measures are known for producing an oil flow which reduces the thermal strain of the converter lock-up clutch.

The oil stream produced through the known measures however causes the torque which can be transferred by the lock-up clutch to be reduced as a result of the dynamic or kinetic processes which occur in the oil stream. The torque transfer capacity of the lock-up clutch thereby decreases with increasing speed and with increasing volume flow of oil. In the case of torque converters which are to be completely bridged from a certain speed the system pressure must be mate correspondingly high so that the component parts, more particularly the piston, also have to be strengtheneo and a more powerful pump is required.

Furthermore the volume flow of oil is again increased through the higher pressure whereby additicnal losses occur.

The said reduction in the torque transfer capacity of the converter -ock-up clutch is due inter alia to the forces which are produced by the dynamic processes and which act on the oil flowing radially inwards thereby causing a pressure increase in this oil. These forces produce an axial component which acts on the piston in the opening direction of the clutch.

A further disadvantage of the measures known up until now lies in the fact that the oil stream is very much dependent on the terserature or viscosity of the oil and on the pressure difference between the pressures arising either side of the converter piston. With the solutions according to US-PS 4 369 543 the flow resistance produced through the channels mst thus be designed for the critical case, this means that even at the maximum oil temperature possible only so much oi is allowed to flow away through the channels so that the s::-stem pressure in the torque converter does not collapse to an unacceptably low level. Even with the solutions according to US-PS 4 969 543 the oil stream flowing through the channels is dependent on the pressure difference between the two chambers. This pressure difference Is the parameter for the torque of the clutch and can thus not be used to adjust a desired volume flow. In order to restrict the losses in the torque converter to an acceptable level the volume flow of oil which exists at maximum pressure difference, thus at maximum coupling torque must be limited. The cooling oil stream can indeed be made sufficiently large for the maximum coupling torcue, but for average and lower torques the volume stream present is then too small for many cases as a result of the Low pressure difference.

The object of the present invention is to improve the lockup clutch employed in hydrodynamic torque converters known up until now, more particularly by increasing the torque transfer capacity and reducing the thermal strain, more particularly in the area of the friction faces of the lockup clutch. Furthermore the thermal strain of the oil is also to be reduced. A further aim of the invention is to optimise the cooling oil stream through the converter clutch throughout the entire operating range of the torque converter, as well as to improve the heat exchange in the area of the friction faces of the lock-up clutch between the oil and the adjoining component parts. ohe measures according to the invention are also to allow a better control and regulation of the torque transferable by the lock-up clutch and of the slip present on the lock-up clutch so that the torque shocks or torque irregularitIes occurring in the drive train or the internal combustion engine can be better damped through the slip whereby the comfort is increased.

According to the invention there is provided a friction clutch for a torque converter, said clutch having at least one pair of annular interfaces that are adapted to engage while rotating in one direction and continuously slipping relative to each other to effect torque transmission in bypass relationship to torque transmitted by hydraulic fluid in said converter, one of said interfaces having at least one circuitous groove, said groove extending a substantial annular distance about said one interface and adapted to direct hydraulic fluid in said torque converter from a high pressure cavity at an outer edge of said one Interface and across the width of said one interface in a circuitous path in at least three traversing passes to a low pressure cavity at an inner edge of said one interface, each c said passes extending substantially the entire width of said one interface, said groove having an inlet section that defines one of said passes and has an inlet that is adapted to open said groove at said outer edge of said one interface to said high pressure cavity, and said groove having an outlet section that defines another of said passes and has an outlet that is adapted to open said groove at said inner edge of said one interface to said low pressure cavity at an angular location remote from said inlet.

As a result, the dynamic forces acting on the oil as the converter rotates as a result of the radial flow of the oil can be absorbed at least substantially and at least in the axial direction. It can thereby be achieved that practically no axial force components produced through the resulting dynamic forces and which would cause a reduction in the torque transferable by the lock-up clutch can act on the piston. Thus through the design according to the invention the dynamic forces acting on the oil and producing an increase in the oil pressure are at least partially neutralised with regard to their effect on the clutch closing force acting on the piston.

By guide channel is meant both the radially allied channels with a closed cross-section per se and also cIrcular ringshaped radially inwardly aligned chambers. A cIrcular ringshaped chamber of this kind can also be divided into several radially aligned channels wherein these can nave a closed cross-section per se. As a guide channel can also be used a tubular component which is connected with at least one output side of a channel provided in the radial area of the friction faces. Thus the oil flowing out from at least one such channel can be guided radially inwards through a pipe wherein the outflow side of this pipe can again open into the first chamber or however opens into a special oil return guide. An oil return guide of this kind can be termed for example by channels which are provided in the driven hub of the turbine and in the longitudinal direction of the gearbox input shaft. The tubular guide for the oil flowing radially inwards when the converter lock-up clutch is closed can be provided in the first chamber or in the second chamber.

It can be particularly advantageous for the construction of the torque converter if the guide channel is supported by a component such as eg a housing or ring piston, having friction faces. Furthermore it is expedient if te channels opening into a guide channel allow an oil flow from the second chamber into the first chamber. To this end corresponding openings or ports can be provided in the component parts adjoining the friction faces or in the component parts forming the friction faces. These openings or ports can thereby be designed so that they act like a throttle or similar to a nozzle. In order to ensure optimum support of the pressure increase which occurs as p result of the dynamic processes in the oil stream it is particularly advantageous if the walls defining the guide channel are fixed axially relative to each other and thus are connected together at least practically rigid. In order to form the guide channel or chide chamber it is possible In a simple way to provide a radial oil guide wall on one of the component parts forming the one chamber wherein tis wall is fixed axially with the component part and together with the radial areas of the latter defines at least one radially extending space which forms the guide channels 1 or guide chamber and in which the oil stream is directed radially inwards. The walls defining the guide channel can thereby be connected advantageously rigidly with a component part supporting one o the friction faces.

A particularly simple construction is produced through the arrangement of the first chamber axially between the ring piston and a radial wall of the housing. It can thereby be expedient if the ring piston is provided axially between the radial wall of the housing and the turbine wheel. It can furthermore be advantageous if the guide channel or the oil guide wall defining same is supported by the piston.

The oil guide wall defining the guide channel or guide chamber can advantageously be housed in the fIrst chamber wherein this wall can be designed and arranged so that the first chamber is divided into two sub-chambers. In many cases it can however also be expedient if the ol guide wall is provided on the side of the ring piston remote from a chamber, thus on the side of the ring piston facing the turbine wheel.

At least the turbine wheel and the pump wheel and if required the guide disc can be housed in the second chamber.

A particularly advantageous construction can be produced if at least one of the component parts supporting the friction faces has a friction lining. This friction lIning can be supported by the piston or by a radially aligned wall section of the housing. Fixing the friction lining on the corresponding component part can be achieved for example by adhesive. When using a friction lining it can oe expedient if the channels are defined at least in part directly by the friction lining. To this end the channels can be provided in the friction lining. The channels can be formed at least in part eg by indentations or cut-out section: or grooves formed in at least one friction lining. Through such a design it is possible to produce a continuous oil stream between the two chambers which is directed over the friction lining. A particularly good heat exchange is thereby produced between the friction faces and the oil so that viewed overall a low or lower thermal strain of both the component parts forming the friction faces and of the oil is achieved.

It can be particularly expedient if the inlet of the channels provided in the radial area of the friction faces lies' radially further out than the outlet and it these channels open into a guide channel or'guide chamber. The guide channel or guide chamber can extend at least over 50% of the radial extent of the first chamber. The greater this radial extension then the smaller are the resetting forces acting on the piston of the lock-up clutch and produced as a result of the dynamic processes acting on the oi.

In order to feed the channels with oil it can be advantageous if the relevant inlet side of the cha.--nels is connected with an axial supply opening provided in the ring piston and/or in the oil supply wall. The outlet side of the channels can be connected with an axial port provided in the ring piston and/or in the oil guide wall and itself opening into a guide channel. In the case of channels which are formed by indentations or cut-out sections in the friction lining and/or in the area of a friction face of a component part, the relevant outlet side of the channels can also be formed so that it opens directly into the guide channel.

In order to achieve a particularly effective heat exchange the channels provided in the radial area of the friction faces are designed zig-zag or meander-shaped. The largest possible guide length for the oil can thereby be achieved within the area of the friction faces. The length of the channels as well as their cross-section must thereby be adapted to the desired volume flow of oil. To this end it is expedient if te grooves or cut-out sections forming the channels are corparatively deep wherein on creating these channels in a friction lining this depth can extend preferably over practically the entire depth of the friction lining. It is thereby particularly advantaceous if the channels are formed by oblong punched-out areas in the lining. The relevant inlet side of the channels can advantageously be provided in the outer edge area of the friction lining and the relevant outlet side in the radially inner edge area of the friction lining. For the heat exchange between the friction faces and the oil it can furthermore be expedient if the zig-zag or meander shaped channels run in the circumferential direction of the friction lining so that the oil - seen over the radial width of the friction faces - is guided repeatedly radially to and fro. The channels can have for this purpose at least two deflections wherein grooves with at least four deflect ions have proved particularly advantageous.

When using a loc-up clutch with conical friction faces it can be particularly advantageous if the friction lining is made as a cone wherein the conical shape Is formed by joining together the two ends of the cone winding. The ring-like friction lining can however also ne made from several segment or crescent-shaped friction elements which can be combined together to produce a ring-shaped or frustoconical shaped design. By using friction linen segments it is possible to reduce the amount of material used since the amount of waste resulting can be significantly reduced. It can be particularly expedient if the starting material for the friction linIngs is coated with adhesive foil prior to punching since tis ensures easier handling of the finished linings.

When the channels or grooves are punched through in the friction lining it is expedient if the latter has a continuous contour both radially outwards and radially inwards since it can thereby be ensured that handling is practically free of deformation and thus even when the friction linings are stuck on the support such as eg the piston no deformations can occur which would impair the function of the lock-up clutch.

The oil stream flowing through the channels can be adjusted by at least one valve in dependence on at least one operating parameter of the torque converter and/or of the machine driving same and/or of the gearbox driven by the torque converter. A parameter of this kind can be formed for example by the oil temperature, by the drive speed of the machine or the output speed of the torque converter and input speed of the gearbox respectively. It is also possible to use as a parameter in a particularly advantageous way the pressure difference between the two chambers. It can thereby be particularly expedient if the valve has a regulating characteristic which when the lock-up clutch is closed ensures a practically constant oil flow, thus a constant volume flow, over the entire operating range of the torque converter. In many cases it can however also be expedient to provide other characteristic lines for the volume flow, more particularly those which are dependent on the slip of the lock-up clutch or on the amount of heat arising as a result of the slip. A particularly simple construction can be produced where the valve is designed as a volume flow valve which adjusts the desired volume flow in dependence on the pressure difference between the two chambers.

The channels provided for conveying cooling oil and channels can be connected to at least one valve by means of which the volume of oil which can be conveyed through the channels can be adjusted by changing the through-flow cross-section in dependence on at least one operating parameter.

The valves can advantageously be provided on the inlet side and/or outlet side of the corresponding channels or guide channel. A valve of this kind can however also be mounted in the area between the inlet side and the outlet side of a channel or a guide channel.

According to a further embodiment of the channels these can also be formed by moulded areas such as eg grooves provided in the material of at least one of the component parts, such as eg the ring piston and/or housing, supporting or forming the friction faces.

A particularly simple regulation of the volute flow of cooling oil can be achieved where the through flow crosssection of a valve is variable in dependence on the pressure difference between the two chambers, wherein it can be expedient if the through flow cross-section o the valve becomes smaller as the pressure difference between the two chambers increases. Such valves can advantaeously be designed and arranged so that they have with regard to their function practically no or only a slight dependence on the centrifugal force acting on them. Advantageously the valves can also have a volume flow characteristic which is not proportional to the root of the pressure difference between the two chambers.

The oil flow regulating valves can also be designed as solenoids.

The ratio between the thickness of the friction lining and the average depth of the grooves viewed over the longitudinal extension of the grooves can advantageously be in the order of between 2.7 and 1.3. The death of the grooves can thereby be in the order of between '.2 and 0.8 mm, preferably between 0.3 and 0.6 mm. The grooves can have practically the same depth over their entire extension. In many cases it can however also be expedient if, viewed over the longitudinal extension of the grooves, the depth varies.

Furthermore the grooves can have at least substantially the same width over their longitudinal extension. Hover grooves can also be expedient which change in width over their longitudinal extension.

The throttling of the cooling oil stream can thus be carried out in the flat zig-zag grooves running substantially over the entire width of the lining. Between their deflections the grooves can have substantially straight parts of about 10 to 40 mm in length. The groove width can be in the order of between 3 and 10 mm. In order to ensure a through flow of a maximum of about 10 litres per minute which is the most favourable in many cases with a pressure of about 5 bar arising on the side of the piston facing the turbine, the most expedient groove depths are in the order of 0.3 mm.

The number of individual zig-zag or meander-shaped grooves running from the radially outer area to the radially inner area of the friction lining is preferably about 4 or 12 which are spread at least approximately uniformly round the circumference of the friction lining. The division or spacing between two deflect ions of the grooves on the same radial side of the friction lining in relation to the radial overall height of the meander-shaped guide groove is preferably in the order of between 0.6 to 1.3, preferably between 0.8 and 1.1.

In order to improve the cooling effect pocket-like recesses or indentations can be formed in the radially outer and/or in the radially inner edge area. These pocket-like moulded areas which are open radially outwards or radially inwards can have approximately the same height as the cooling grooves. The moulded areas can however extend over the full lining height so that they can be made in a simple way during manufacture of the lining or by cutting out. These pocket-like moulded areas can each advantageously be mounted - viewed circumferentially - between two adjoining deflections.

In order to discharge the amount of heat arising in the area of a friction face engagement of a converter lock-up clutch (lock-up), friction linings have proved most suitable where the ratio of the surface taken up by the grooves and/or pocket-like moulded areas to the remaining lining riot ion surface is in the order of about 0.7 to 1.8, preferably in the order of 1 and 1.5.

By providing pocket-like moulded areas in the radially outer and radially inner edge areas of the friction lining it is possible to improve or increase the cooling action in the area of the friction engagement of the lock-up clutch by utilising the drag current. It can be particularly advantageous if, viewed in the rotary direction of the friction lining, the radially outer inlet side of a cooling groove trails the outlet side of this cooling groove which is provided radially inwards. It can be advantageous if the through-flow direction, viewed in the circumferential direction of the friction lining, through the grooves agrees with the direction of the drag currents. In the case of a converter construction where the friction linking is supported by a component part, such as a piston, rotatable with the turbine wheel, this drag current is produced through the component part, such as in particular the converter housing, which forms a counter friction ace for the friction lining. In the event of slip occurring in the lock-up clutch of the converter, the converter 'housing, provided torque is being transferred from the drive motor to an output component part, has a greater rotational speed than the friction lining provided with grooves so that the cooling medium flowing through the grooves is accelerated through the converter housing and the faster rotating counter friction face. Through the arrangement and guide of the grooves the speed influence which otherwise cannot be avoided can be substantially reduced to the torque transferable by the clutch or substantially eliminated.

The friction lining can also be fixed, eg by sticking, on the pump side or on a surface formed by the converter housing. The piston of the converter lock-up clutch then only has a metal counter friction surface.

Furthermore the formation of the grooves can also be produced directly in the material forming the piston and/or the converter housing. With a design of this kind it is also possible to use a friction lining without grooves.

However friction linings with grooves can also be used wherein these grooves can then also be designed differently from those described in the present application.

It can be particularly expedient if the grooves are designed so that viewed over the length of the grooves and for the pressures which occur on the input side and output side of the grooves when driving a vehicle equipped with a torque converter a substantially turbulent current is provided in the area of the grooves. The grooves can thus be designed so that the pressure difference provided between the input side and the output side of these grooves produces a turbulent current in the grooves. The formation of a turbulent current within the grooves can be positively affected by a corresponding guide and configuration of the grooves.

The invention will now be explained in detail with reference to Figures 1 to 12 in which: Figure 1 shows a sectional view through a device according to the invention; Figure 2 is a diagram with torque characteristic lines of lock-up clutches; Figures 3 to 5 show further design possibilities according to the invention of a torque converter with lock-up clutch; Figures 6,7 and 12 show design possibilities of a friction lining for a lock-up clutch according to the invention; Figure 8 shows a detail of a lock-up clutch according to the invention; Figures 8a to 11 show valves for regulating the volume flow of oil as well as details of lock up clutches equipped with valves of this kind.

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

The housing 2 is formed by a housing shell 4 adjoining the driving shaft or internal combustion engine, as well as by a further housing shell 5 fixed on the first shell. The two housing shells 4 and 5 are connected together fixed and sealed radially on the outside 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 f the outer sleeve-like area 4a of the housing shell 4 Axially between the pump wheel 7 and the radial wall 9 of the housing 4 is a turbine wheel 10 which is connected fixed or rotationally rigid with a driven hub 11 which can be coupled rotationally secured with a gearbox input shaft through internal gearing. A guide disc 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 in the housing of a gearbox. In the inner space 14 formed by the two housing shells 4,5 is a lock-up clutch 15 which is mounted operationally parallel with the torque converter 3. The lock-up clutch 15 allows a torque coupling between the output hub 11 and the driving housing shell 4.

A rotationally elastic damper 16 which in the illustrated embodiment is provided between the ring-like piston 17 of the lock-up clutch 15 and the driven hub 11 is connected actively in series with the lock-up clutch 15. The rotationally elastic damper 16 comprises in known way energy accumulators in the form of coil springs. The ring-like piston 17 provided axially between the radially aligned wall 9 and the turbine wheel 10 is mounted for restricted axial displacement radially inwards on the driven hub 11. 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 lock-up clutch 15 axially between the ring-like piston 17 and the radial housing wall 9, and a second chamber 20 which contains inter alia the pump wheel 7, the turbine wheel 10 and the guide disc 12.

The housing shell 4 forms with radially outer areas a conical friction face 21 whose imaginary cone tip is directed axially away from the turbine wheel 10. The conical friction face 21 can be brought into friction engagement with a friction lining 22 which is supported by the conical area 23 of a support plate 24. The support plate 24 is in turn supported again by the ring-like piston 17 which is likewise deep drawn from sheet metal plate.

With more recent concepts for a drive train, eg of a motor vehicle, the lock-up clutch is operated with slip t'r-ough at least a large part of the operating area of the torque converter wherein during the slip phases in the friction engagement area 19 a loss of power in the form of heat takes place which under certain operating conditions can be very high amounting to several kilowatts. Such operating conditions can exist for example when driving uphill with trailers and when changing from the non-bridged to the practically bridged stage of the converter clutch. Such concepts for operating a converter lock-up clutch with slip have been proposed for example by German Patent Applications P 42 28 137.7-12 and P 42 35 070.0-12.

In order to avoid unacceptably high temperatures in the friction engagement area 19, and thus to ccnteract destruction of at least the friction lining surface and 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 25 formed in the friction lining 22 and through which when the lock-up clutch 15 is practically closed a permanent oil stream can take place between the second chamber 20 and the first chamber 18. The oil stream is thereby directed over the friction lining 22 and the friction surface 21. The oil channels 25 are provided with the optimum shape so that a good heat exchange can take place between the component parts causing the friction engagement in the area 19 and the oil flowing through. A preferred shaping of the channels 25 will be described in further detail in connection with Figures 6 and 7.

The oil inlet end of the channels 25 lying radially further outwards is connected with the second chamber 2C through ports or bores 26 in the piston 17 and in the support plate 24. The outlet end of the channels 25 lying radially further inwards is connected with the first chamber 18.

The support plate 24 fixedly connected axially to the piston 17 forms in connection with the radial areas of the piston 17 a sub-chamber 18a which serves as a guide channel for the oil which when the lock-up clutch 15 is closed flows through the channels 25 radially in the direction of the rotary axis 27. The guide channel 18a is in communication with the radially inner end areas of the channels 25 through openings or bores 28 provided in the support plate 24. Radially inside, the support plate 24 has axial indentations 29 which serve as spacers between the support plate 24 and the ring piston 17. Between the indentations 29 the guide channel 18a which in practice is designed as a ring-like chamber opens radially inwards whereby a connection is made with the sub-chamber 18b formed between the radial housing areas 9 and the support plate 24. In the illustrated embodiment the piston plate 17 and the support plate 24 are connected together by rivets in the area of the indentations 29.

Radially outside the guide channel 18a the ring piston 17 has a ring-like axial indentation 30 which likewise serves as a spacer between the support plate 24 and the remaining areas of the ring piston 17. The indentation 30 furthermore increases the deformation resistance of the support plate 24 in the friction engagement area 19. In the area of the ring-like indentation 30 there is a radial seal between the two component parts 17 and 24. The cooling oil stream takes place when the lock-up clutch 15 is closed starting from the second chamber 20 through the ports 26, the oil channels 25, the openings 28 and the guide channel 18a radially inwards into the area of the driven hub 11. This cooling oil stream can then be discharged in the area of the driven hub 11, for example through a hollow shaft or through a channel provided for this purpose, namely preferably first into an oil cooler. From this oil cooler the oil can then be returned to a sump and from there back into the hydraulic regulating or control circuit.

The component parts 17 and 24 forming the guide channel 18a are supported axially relative to each other or connected together so that the oil flowing radially inwards therein can exert no axial force components on the axially displaceable piston 17 which would cause a change, more particularly a reduction, in the torque transfer capacity of the lock-up clutch 15. This is achieved in that the forces arising in the oil as a result of the radially inwardly aligned oil stream, or the pressure increases which create an axial force component on the piston 17 and on the support plate 24, are axially absorbed so that a closed force flow is produced. This is ensured with the illustrated embodiment according to Figure 1 through the axial support of the metal plate 24 on the piston 17.

The said forces which are due to the dynamic processes in the oil, and thus also the pressure increases caused by same reduce very considerably, in the case of torque converter constructions with lock-up clutch, as known for example from US-PS 4 969 543, the maximum torque transferable by the lock-up clutch as the speed increases. With the known torque converters with cooling oil stream these torque losses which are due to dynamic processes occur when the oil stream flows from radially outwards to radially inwards between the radial housing wall and the piston of the practically closed lock-up clutch wherein these losses become greater as the volume flow of oil increases. One cause for the decrease in the torque transfer capacity of the lock-up clutch with increasing speed or increasing volume flow of oil is probably due to the Coriolis acceleration acting on the oil as the oil flows from radially outwards to radially inwards and which as a result of the rotation of the torque converter acts in the rotary direction on the oil and can cause an increase in pressure in the oil flowing radially towards the axis of rotation 27.

With the present invention the pressure increases caused as a result of the radial oil flow and the axial forces thus produced on the component parts guiding the oil flow are axially absorbed so that they have practically no effect or at least only a significantly small effect on the closing force of the lock-up clutch 15 and thus on the torque transferable by same. These undesired pressures or forces are axially absorbed according to the invention.

With the embodiment shown in Figure 1 the support plate 24 or the guide channel 18a runs radially up to the driven hub 11, thus comparatively far inwards. In many cases however it can also be expedient if the guide channel 18a only extends over a partial area of the radial extension of the ring-like piston 17 so that with increasing speed or increasing volume flow a specific change in the torque transferable by the lock-up clutch 15 takes place. In most cases however it is expedient if the guide channel 18a runs over at least 50% of the radial extent of the ring-like piston 17. Furthermore it is possible to guide only a part of the oil stream through the guide channel 18a and to guide the rest radially inwards through the sub-chamber 18b. To this end the support plate 24 can have connecting openings between the sub-chamber 18b and the guide channel 18a.

These connecting openings can be designed accordingly for the desired effect and are arranged in a specific radial distance from the axis of rotation 27.

In the diagram according to Figure 2 the speed of the torque converter or housing 2 is shown on the abscissa and the ratio of the torque transferable by the lock-up clutch 15 to the pressure difference present either side of the converter piston 17 is shown on the ordinate. The line 31 shows the torque path over the speed for a given, thus constant differential pressure P at the piston of a conventional converter lock-up clutch, thus of a lock-up clutch wherein with a closed clutch no oil stream exists from one side of the piston to the other. It is evident that the transferable torque remains at least substantially constant with such a lock-up clutch over the speed at a given P.

Torque converters with a lock-up clutch of this kind are shown by way of example in US-PS 4 646 763.

The line 32 represents a possible path of the torque transferable by the lock-up clutch 15 over the speed at a given P of a hydrodynamic torque converter wherein an oil stream is present from the second chamber 20 into the first chamber 18. Such hydrodynamic torque converters are known for example through US-PS 4 445 599 and US-PS 5 056 631.

With these designs channels or openings are provided in the area of the friction linings and/or the piston of the lockup clutch and allow an oil flow from the second chamber which holds at least the pump wheel and the turbine wheel, into the first chamber which is defined by a radial wall of the housing and the piston. This oil flow or oil stream leads to a reduction in the maximum torque theoretically transferable for a given P by the converter lock-up clutch - as a result of the current losses in the supply and return line and on flowing through the converter. This can be seen from Figure 2, namely at low speeds for an identical P the maximum torque transferable according to line 32 is less than the torque of line 31 corresponding to the same speed.

The static losses are still superimposed by the dynamic losses which likewise reduce the torque transferable by the lock-up clutch. These are caused by the oil stream directed radially inwards in the first chamber. As can be seen from the path of the line 32 the torque transferable for a given P by the lock-up clutch is considerably reduced with increasing speed through the dynamic losses.

Through the support according to the invention of the forces or pressure increases in the chamber 18 occurring as a result of the radially inwardly directed oil stream, it is reached that for a given P the torque transferable by the lock-up clutch over the speed of the torque converter does not drop according to line 32 but remains at least substantially practically constant according to the dotted line 33. Depending on the desired behaviour of the lock-up clutch the line 33 can also have a different path. Thus if required viewed over the speed, a certain drop of the transferable torque can also be provided. With the optimum design however of the at least one guide channel 18a for the oil stream it can be achieved that practically only static losses occur compared to the ideal case in terms of torque transfer corresponding to line 31.

The above view is idealized since the friction in the fluid and the friction between the fluid and the guide walls was not taken into account.

The embodiment according to Figure 3 differs from that according to Figure 1 in that the component part 117 does not run radially outwards up to the outer contour of the sheet metal plate 124 and is not guided axially displaceable radially inwards on the output hub 11. The metal plate 124 is in the same way as the ring piston 17 according to Figure 1 set radially centred on the output hub 11 and displaceable axially to a restricted amount. In the embodiment according to Figure 3 the sheet metal plate 124 thus in practice forms the piston of the lock-up clutch 115 and the ring-like component 117 forms a reinforcement for the plate 124. With an embodiment according to Figure 4 the component part defining the guide channel 218a is provided in the form of a ring-like metal plate 224 on the side of the ring-like piston 217 remote from the first chamber 218. The piston 217 has in the radially outer area of the guide channel 218a ports 228 which communicate with the channels 225 in the friction lining 222. Drain openings 234 are formed in the piston 217 in the radially inner area of the guide channel 218a and open into the first chamber 218.

As already mentioned, the guide channels 18a, 118a, 218a can in practice be designed as ring-like chambers. However several radially aligned channels can also be provided which each communicate with at least one of the supply openings 28, 228. Thus for example instead of guide plates 24, 224 it is also possible to use individual small tubes which communicate radially outwards with one of the openings 28, 228 and extend radially inwards towards the axis of rotation. Also it is not necessary to direct the radially inwardly aligned oil stream back into the first chamber 218 or sub-chamber 18b but the outflow for the radially inwardly directed oil flow can also be provided through a return line provided particularly for this purpose. Thus for example this oil flow could also be evacuated through at least one radial bore formed in the output hub 11 and opening into a drain channel.

With the hydrodynamic torque converter 303 shown in Figure 5 a first chamber 318 and a second chamber 320 are again provided which are separated from each other through the piston 317 of the lock-up clutch 315. The ring piston 317 is centred axially displaceable on the output hub 311. The rotationally elastic damper 316 is mounted in the torque flow between the output hub 311 and the ring piston 317.

When the lock-up clutch 315 is closed an oil stream is provided between the second chamber 320 and the first chamber 318 and is directed in the chamber 318 radially inwards. To this end at least one inlet opening 326 is provided on the piston 317. In the illustrated embodiment this inlet opening is formed by a nozzle 326a supported by the piston 317. The oil flowing through the opening 326 is guided in grooves or channels 325 in the friction engagement area 319 between the friction face 321 of the housing 302 and the friction face of the friction lining 322. On the outlet side of the grooves or channels 325 the oil then enters the first chamber 318 and flows radially inwards.

The oil flow is thereby guided radially inwards between the housing wall 309 and a radial ring-like support plate 324.

The plate 324 is fixed on the output hub 311 and is supported axially, thus in the axial direction towards the piston 317. This support has the effect that the pressure increases occurring in the radially inwardly flowing oil cannot in practice act on the piston 317 since the axial forces produced by these pressure increases are absorbed by the plate 324.

Instead of being connected to the output hub 311, the oil guide plate or support plate 324 could also be connected axially to the housing shell 304. A closed force flow in relation to the axial forces arising would also then exist.

Instead of being formed in the friction lining 25, 225, 325, the oil grooves or channels 25, 225, 325 could also at least in part be provided in the adjoining housing wall and/or in the piston of the lock-up clutch and/or eg in the embodiments according to Figures 1 and 3 also in the support plate 24, 124. A piston with oil guide grooves is known for example through US-PS 5 056 631.

The guide of the oil flowing between the second chamber and the first chamber radially inwards and the associated support or compensation of the impulse forces or pressure increases arising in the oil is not restricted to embodiments where the oil flow takes place directly in the area of the friction faces of the converter lock- clutch, but can also be used in those designs as known for example from US-PS 4 493 405 and US-PS 4 445 599.

Figure 6 shows a friction lining 422 which can be used with a converter lock-up clutch according to Figures I and 3 to 5. The friction lining 422 has a radially outer cohesive area 422a extending over the circumference and a likewise cohesive radially inner area 422b extending in the circumferential direction. Cut-out sections 43 having a zig-zag or meander-shaped path are formed in the central area 422c between the outer and inner area 422a, 422b. The cut-out sections in the illustrated embodiment are guided meander-shape or sinusoidal in the circumferential direction of the friction lining 422. Through a guide of the cut-out sections 435 of this kind it is possible to ensure a particularly long channel length for the oil passing through whereby a good heat exchange is achieved between the oil flowing through and the component parts forming the friction faces of the lock-up clutch. A low thermal strain both on the friction faces or the component parts forming same and on the oil present in the area of the friction faces can now be achieved.

The length measurements and shaping of the channels or punched-out areas 435 must be carried out so that the flow resistance appearing therein is designed for the critical operating state of the torque converter or lock-up clutch, this means that even with the maximum possible oil temperature only so much oil is allowed to pass from the second chamber to the first chamber so that the system pressure in the torque converter does not break down. It is expedient if the cooling oil stream guided through the punched-out areas or grooves 435 has the slowest possible dependence on the oil temperature.

The friction lining 422 has in the illustrated embodiment nine channels or grooves 435 distributed evenly over the circumference. It is expedient if at least three such zigzag channels or grooves are formed in the friction lining 422.

The friction lining 422 illustrated in Figure 6 is shown in plan view and is stuck onto the frusto-conically shaped areas of the corresponding ring piston or of the corresponding support plate or of the correspondIng housing shell. The friction lining 422 is made as a cone winding so that when fitting together the two end areas 436, 437 a conical or frusto-conical shape is produced. A particularly small amount of material can be used by dividing the friction lining 422 into several sector-like carts 438.

This is shown in FIgure 7. The friction lining segments 438 are then stuck on the corresponding support part. It can be particularly advantageous for this if the lining material or the starting blank is coated with an adhesive foil on one side at least befcre cutting out the recesses 435. Easier handling is thereby ensured. This handling becomes further improved if the lining segments 438 or the lining 422 has(ve) radially cutwards and radially inwards a cohesive continuous area.

With the embodiment according to Figures 6 and 7 the grooves or punched-out areas 435 are closed radially to the outside and radially to tne inside. This is possible by suitably designing the component parts clamping the friction lining.

The design must thereby made so that at the radial outer end 439 of the channels 435 the oil can flow into te channels 435 and can flow ot again at the radially inner end 440. To this end in the er.nodiment according to Figures 1 and 3 to 5 corresponding bores or recesses are provided in the adjoining parts which are in communication wit the corresponding ends of the channels 435. Instead of bores or cut-out sections the corresponding parts could however also have indentations or impressions which ensure a connection of the corresponding end 439 and/or 440 wit the corresponding chamber eg 320 and/or 318. Figure 5 shows in chain-dotted line a corresponding connection which is formed by an indentation 441 in the piston 317. This moulded area or indentation is off-set circumferentially relative to the oil inlet side 326 and is connected with a radially inner end of a channel 325.

It is expedient if the channels 435 have, viewed over their length at least two deflections, thus at least three arms or two curves. In the embodiment according to Figure 6 the channels have six deflections and thus seven arms. With a sinusoidal or snake like guide of the channels 435 these would have six adjoining curves.

Figure 8 shows a further possibility for making a connection between a radially inner end 440 of a friction lining 422 and the first chamber 18 formed between the housing 2 and the ring piston 17. This connection is ensured through an axial step 2a on the housing shell or on the cover 4. This step 2a is thereby arranged so that it extends radially outwards over the end area 440 of the corresponding channel 435. This step 2a can also be formed ring-like, thus extend over the entire circumference so that then all the end areas 440 of the channels 435 have a connection with the first chamber 18.

In order to reduce the dependence of the oil volume flowing between the second chamber and the first chamber on the oil temperature or oil viscosity and on the pressure difference existing between the second chamber and the first camber, a way can be provided to adjust the volume flow in dependence on the oil temperature or oil viscosity and/or pressure difference between the pressures arising either side of the piston.

Such means are shown in the form of a valve 542 in Figures 8a to 10.

In Figure 8a the valve 542 is supported by the piston 517.

The valve 542 has a housing 543 which is fixed on the piston on the side of said piston 517 remote from the friction lining 522. In the illustrated embodiment the housing 543 has for this purpose a ring-like outer shoulder 544 which is set in a bore 545 of the ring piston 517 where it is fixed eg by force fit.

As can be seen from Figure 9, an axially displaceable piston 546 is housed in the housing 543 of the valve 542. The piston 546 has an axial attachment 547 which is axially displaceable in a recess 548 which opens to the outside. By axially displacing the piston 546 it is possible to change the cross-section of the outlet 549 which is set between the recess 548 and the axial attachment 547. This change in cross-section is thereby carried out by correspondingly shaping the axial attachment 547 and/or the recess 548. The recess 548 is in the illustrated embodiment formed by a ring-like socket 550 of L-shaped cross-section and fixed on the housing 543 so that its axial attachment 551 projects into the cylinder space 552. A graded spring 553 set on the axial attachment 551 biases the piston 546 towards the housing base 554. It is thereby ensured that with a low differential pressure between the two chambers, eg 20 and 28 or between the chamber 20 and the guide channel 18a a comparatively large oil stream can flow through the valve.

The valve 542 has inlets 555 which produce a connection between the second chamber 20 (Figure 1) and the cylinder space 552. The recess 548 forming the outlet of the valve 542 opens or is in communication with at least one oil guide channel 535 provided in the friction engagement area 519.

This connection is thereby designed so that the oil flowing through the valve 542 is preferably guided from one end of a channel 535 to the other end of this channel and from there flows out radially towards the rotary axis of the corresponding torque converter.

In the illustrated embodiment the axial attachment 547 of the piston 546 has a groove 556 which has a geometry or path such that as the axial displacement of the piston 546 increases to the left so the outlet cross-section 549 becomes smaller. Through a corresponding design of the groove 546 and of the force-path characteristic of the spring 553 it can be achieved that a practically constant volume stream flows through the valve 542 over the entire useful area of the corresponding torque converter. This has the result that a volume stream is provided which is practically independent of the difference of the pressures arising either side of the piston of the lock-up clutch. If necessary however different characteristic lines for the volume stream can also be produced through a corresponding design of the attachment 547, recess 548 and energy accumulator 553. Thus for example through a correspondingly designed valve 542 the adjusting volume flow of oil can at least increase slightly or at least decrease slightly with increasing pressure difference. If necessary the volume flow can also be completely interrupted from a predetermined pressure difference between the two chambers. The flow regulator valve 542 is however preferably designed so that the volume flow is kept practically constant and is thus practically independent of fluctuations in the supply or load pressure arising at the inlets 555. A flow regulator valve of this kind furthermore has the advantage that this can be designed so that temperature changes in the oil can be substantially compensated which means that the volume flow is substantially independent of the temperature changes in the oil.

The volume flow regulator valve 542 is mounted on the inlet side or at the beginning of the cooling current flow.

The flow regulator valve 642 illustrated in Figure 9a can be used in place of the valve 542, wherein the holder for the valve body on the piston must be modified accordingly, or however in place of the nozzle 326a. The valve 642 has a housing 643 which forms a cylinder chamber 652 for the piston 646. The open side of the housing 643 is partially closed by a disc 650 which has an opening 650a. Between the housing base 654 and the piston 646 is a calibrated spring 653. The spring 653 is at least partially housed in an axial indentation 646a of the piston 646. The cylinder space 652 is divided by the piston 646 into two spaces wherein the right space 652a is supplied through the opening 650a with oil which has a pressure which corresponds to the pressure in the second chamber which holds at least the pump and turbine wheels. The left space 652b is supplied through a bore 657 provided between the two spaces 652a and 652b.

The bore 657 serves as a means for setting a pressure drop P between the two spaces 652a and 652b. A regulator aperture 658 is connected in series with the bore 657 and regulates the volume flow of oil passing through this regulator aperture 658 in dependence on the pressure arising in the chamber 652a. This is achieved in that the pressure difference P between the two spaces 652a and 652b is set to a specific value by suitably setting the cross-section at the regulator aperture 658. With the invention this value is preferably constant so that a practically constant volume flow is produced. The regulator aperture 658 is formed by radIal openings 648 provided in the housing 643 and whose through flow cross-section can be altered by axially displacing the piston 646. If the pressure in the chamber 652a increases then the piston 646 is moved to the left against the force of the spring 653 whereby the through-flow cross-section of the openings 648 becomes smaller. The pressure present in the chamber 652b is thereby again brought to a higher level, and a pressure difference is again set between the two chambers 652a and 652b which ensures that the desired amount of oil flows through the openings 648.

With the embodiment according to Figure 10 the friction lining 722 is supported by the housing shell 704. In Figure 10 a support plate 704a is provided for this which is fixed on the housing shell or on the cover 704, namely through rivet connections 760. The rivet connections 760 are formed by means of rivet studs which are made out of the material of the housing 704 and engage in corresponding recesses in the plate 704a. The lining support part 704a is set up ccnically radially outwards in a direction axially away from the radial wall 709 of the cover 704 so that a ring-like intermediate space 761 of wedge-shaped cross-section is formed. In the interspace 761 there is the volume flow valve 742 which is supported by the component part 704a.

The interspace 761 has a connection with the second chamber o the flow converter or is part of this second chamber.

The disc-like lining support component 704a can have radially outwards - viewed over the circumference - at least in areas a connection with the radially outer wall of the housing 704 whereby the axial rigidity of the component part 704a is increased. The piston 717 of the lock-up clutch 715 likewise has a conically aligned area 730 which forms a friction face which can be brought into engagement with the lining 722. When the lock-up clutch 715 is closed a defined ccoling oil stream flows through the valve 742 into the channels or grooves 725 provided in the area of the lining 722.

Figure 11 shows a design which allows the individual oil grooves or oil channels 825 provided over the circumference of the lining 822 to be supplied with oil through a single valve :42. For this in the illustrated embodiment a component part in the form of a ring-like cover 862 is fixed on the piston 817 whereby a chamber 863 is formed between this cover 862 and ring piston 817. This chamber extends over the circumference and is connected through openings 826 with each oil inlet end of the individual oil grooves 825.

Depending upon the type of use several valves 842 could also be provIded wherein however the number of valves can be less than the number of individual oil grooves or oil channels 825.

The idea (Figure 11) of supplying several oil grooves or oil channels 825 through the same valve 842 can also be used in a particularly simple way in the design according to Figure 1, namely by providing a corresponding valve in the area of a port 26 in the piston 17. All the channels 25 can thereby be supplied with oil through the ring-like space 17a and through the ports 26 in the component part 24.

It is expedient if the valves 542, 642, 742 or 842 are designed and/or arranged so that the effect of the centrifJgal force acting on same is as low as possible whereby the desired function is guaranteed. This can be ensured by using the lightest possible piston and by arranging the direction of movement of the piston in the axial direction of the torque converter. Through the latter step it is ensured that no components of the centrifugal force acting on the corresponding valve act in the direction of the valve spring. The pistons should be designed as small as possible and made from a light material such as eg plastics or aluminium. In the embodiment according to Figure 1 the sensitivity of the valve 842 to the effects of centrifugal force is reduced by placing this valve 842 additionally on a relatively small radius.

By regulating the cooling oil stream it is possible to set an oil stream at the converter lock-up clutch which is not proportional to the root of the pressure difference between the pressures building up either side of the piston.

In the case of hydrodynamic torque converters as known for example from US-PS 4 969 543 there is the drawback that when the lock-up clutch is closed the volume flow flowing out through same is very much dependent on speed wherein as the speed increases so the volume flow decreases considerably as a result of the previously mentioned dynamic or kinetic processes in the oil. This considerable drawback regarding the functioning of the torque converter can be avoided at least substantially by designing the guide of the volume flow radially inwards. Thus through the design according to the invention for a predetermined or desired system pressure in the torque converter it is possible to set a lower volume flow at lower speeds and thus to use a smaller pump.

Figure 12 shows in part a ring-like friction lining 922 which has meander-shaped or zig-zag shaped grooves 935 which extend in the circumferential direction of the friction lining 922 and which have a similar configuration to the grooves 435 according to Figures 6 and 7. The radially zig zacging or sinusoidal grooves 935 have substantially the same width, viewed over their length, and can advantageously also form over their longitudinal extension at least substantially an identical through-flow cross-section for the cooling oil. The grooves 935 are in the illustrated embodiment open both radially towards the outer edge 922a and radially towards the inner edge 922b and are not closed as with the embodiment according to Figure 6 and 7.

The grooves or channels 935 formed in the friction lining or friction ring 922 can be made during manufacture of the friction lining, thus prior to fixing the friction lining on a support component such as eg a ring piston or plate. The grooves or channels can however also be formed in the friction lining during the fixing eg sticking of this lining on a support component or after fixing has been completed.

Thus quite generally the friction lining eg 922 can first be fixed on the corresponding ring piston and during or after this fixing the channels or grooves can be stamped in the corresponding friction ring or friction lining.

It can be particularly advantageous i the angle touching a radially inner or radially outer deflection 946 of the channels 935, and marked in Figure 12 by reference numeral 945, is in the order of between 30 and 1200, preferably in the order of between 45 and 700. In Figure 12 the angle 945 amounts to about 60". Advantageously the individual grooves 935 distributed over the circumference of the friction lining 922 can be measured so that vieed over the length of the grooves 935 at least a substantially turbulent current is produced. This creates a better heat transfer to the cooling oil. According to a varied embodiment of the invention the turbulent current is achieved by suitably designing the deflection areas 946 along the grooves 935.

The multiple guide of the grooves 935 over the radial width of the friction lining 922 likewise has a positive influence on the cooling action caused by the cooling oil in the friction engagement area of the corresponding clutch. By guiding the grooves or channels according to the invention it is possible to obtain a correspondingly long guide of the cooling medium in the friction engagement area of the clutch whereby a good heat transfer to the cooling medium can be achieved.

In order to produce a turbulent current within the grooves 935 it is necessary to take into account when sizing the grooves the differential pressure arising between the inlet side 939 and the outlet side 940 of te grooves 935. In the case Gf converter lock-up clutches this differential pressure corresponds to the pressure difference between the chambers (18 and 20 in Figure 1) provided either side of the piston of the converter lock-up clutch.

In order to improve the cooling acticn pocket-like recesses or indentations 947, 948 can be formed in the radially outer and/or in the radially inner edge area. These pocket-like moulded areas 947, 948 can be made in a similar way to the grooves 935. In the illustrated e.-.bodiment according to Figure 12 the pocket-like moulded areas 947, 948 can however also have a different shape, such as crescent-shaped or semi-circular. Furthermore the moulded areas 947, 948 can also be designed asymmetrical relative to a radius 949 and viewed circumferentially. In the illustrated embodiment the pocket-like moulded areas 947, 948 - viewed circumferentially - are each provided between two adjoining deflections 946. As the friction lining 922 rotates a current of cooling medium is produced in the pockets 947, 948 and this can be turbulent. As can be seen from Figure 12, the pocket-like moulded areas 947, 948 and the channels 935 cross over one another radially at least in part. As a result of the design and arrangement of the cooling grooves 935 and the cooling pockets 947, 948 the friction surface parts 950 remaining between these cooling devices 935, 947 and 9498 likewise have a zig-zag or meander-shaped configuration.

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 functioning methods described in connection vith the various embodiments.

The present invention is to be considered within the scope and in connection with the prior art listed and this represents additional relevant technical disclosure in respect of he present application.

This application is, indirectly, divided out from application 9413421.0 which describes and claims an hydrodynami torque converter with pur::o. wheel, turbine wheel, guide disc and lock-up clutch with ring piston contained in a housing, wherein either side of the ring piston there is a chamber which can be filled with oil, the ring piston supports at least one friction face 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, channels ere provided in the radial area of the friction faces in at least one of the components forming or supporting the friction faces wherein these channels even with axial contact between the friction faces allow an oil flow from the second of the chambers through the channels radially inwards in the direction of the rotary axis of the torque con-erter, wherein the oil flow after leaving the channels is directed radially inwards within at least one guide channel wherein the component parts defining the guide channel are fixed axially relative to each other in relation to the axial forces acting on same as a result of the oil pressure.

Also divided out from application 9413421.0 are three further di-.-isional applications.

Application 9715218.5 describes and claims an hydrodynamic torque con-erter with a housing containing a pump wheel, turbine w'neel and lock-up clutch with ring piston, wherein either side of the ring piston is an oil-fillable chamber, the ring piston supports at least one friction face which can be brought into friction engagement with a counter friction face, channels are provided in the radial area of the friction faces in at least one of the component parts forming or supporting the friction faces wherein these channels allow even with axial contact between the friction faces an oil flow from one chamber to the other, wherein at least one valve is provided whereby the oil volume which can be conveyed through the channels can be adjusted in dependence on at least one operating parameter of the torque converter and/or the engine driving same and/or the gearing which is driven by the torque converter.

Application 9715216.9 describes and claims an hydrodynamic torque converter with housing containing a pump wheel, a turbine wheel and a lock-up clutch with ring piston wherein either side of the ring piston is an oil-fillable chamber, the ring piston supports at least one friction face which can be brought into friction engagement with a counter friction face wherein radially inside the friction faces the first chamber is formed between the ring piston and a component part supporting the counter friction face, furthermore the friction faces are formed conical and in the radial area of the friction faces channels are provided in at least one of the component parts worming or supporting the friction faces wherein these channels even with axial bearing contact between the friction faces allow a flow of oil fro the second chamber through the channels into the first chamber.

Application 9715217.7 (from which the present application is directly divided) describes and claims a hydrodynamic torque converter with a lock-up clutch with ring piston wherein the ring piston supports at least one friction face which can be brought into friction engagement with a counter friction face wherein at least in the area of the friction face and/or counter friction face channels are provided for conveying a cooling fluid, which channels, viewed in the circumferential direction of the friction faces run zig-zag or meander shape.

Claims (20)

PATENT CLAIMS

1. A friction clutch for a torque converter, said clutch having at least one pair of annular interfaces that are adapted to engage while rotating in one direction and continuously slipping relative to each other to effect torque transmission in bypass relationship to torque transmitted by hydraulic fluid in said converter, one of said interfaces having at least one circuitous groove, said groove extending a substantial annular distance about said one interface and adapted to direct hydraulic fluid in said torque converter from a high pressure cavity at an outer edge of said one interface and across the width of said one interface in a circuitous path in at least three traversing passes to a low pressure cavity at an inner edge of said one interface, each of said passes extending substantially the entire width of said one interface, said groove having an inlet section that defines one of said passes and has an inlet that is adapted to open said groove at said outer edge of said one interface to said high pressure cavity, and said groove having an outlet section that defines another of said passes and has an outlet that is adapted to open said groove at said inner edge of said one interface to said low pressure cavity at an angular location remote from said inlet.

2. A friction clutch as claimed in Claim 1, wherein said inlet section is angled generally opposite to said one direction, and said outlet section is angled generally in the direction of said one direction.

3. A friction clutch as claimed in Claim 1 or Claim 2, wherein a friction material forms said one interface, and said groove is formed in and extends through said friction material.

4. A friction clutch as claimed in any preceding claim, wherein there are a plurality of said grooves angularly spaced about said one interface and connected in parallel with each other between said high and low pressure cavities.

5. A friction clutch as claimed in any preceding claim, wherein the or each groove has a rectangular cross section flow area with a relatively small depth and a relatively large width.

6. A friction clutch as claimed in any preceding claim, wherein the or each groove has a smooth sinusoidal shape with apexes located closely adjacent the edges to which the apexes are nearest.

7. A friction clutch as claimed in any one of Claims 1 to 5, wherein the or each groove has straight traversing sections and pointed return bends joining said straight traversing sections, and said return bends are located closely adjacent the edges to which the return bends are nearest.

8. A friction clutch as claimed in any preceding claim, wherein the or each groove has a uniform cross sectional flow area.

9. A friction clutch as claimed in any preceding claim, wherein there are at least two of said grooves equally angularly spaced about said one interface and there are five of said passes.

10. An annular friction clutch facing for a tcrque converter clutch, said facing having at least one circuitous groove extending a substantial annular distance about said facing adapted to direct hydraulic fluid in a hydrokinetic torque converter from a high pressure cavity at an outer edge of said facing and across the width of said facing in a circuitous pat in at least two traversing passes to a low pressure cavity at an inner edge of said facing, each of said passes extending substantially the entire width of said facing, said groove having an inlet section that defines one of said passes and has an inlet that is adapted to open said groove at said outer edge of said facing to said high pressure cavity, and said groove having an outlet section that defines another of said passes and has an outlet that is adapted to open said groove at said inner edge of said facing to said low pressure cavity at an annular location remote from said inlet.

11. A clutch facing as claimed in Claim 10, wherein said inlet section s angled generally in the direction of rotation of said facing, and said outlet section is angled generally in a direction opposite the direction of rotation of said facing.

12. A clutch facing as claimed in claim 10 or Claim 11, wherein said facing is formed of a friction material, and said groove is formed in and extends through said friction material.

13. A clutch facing as claimed in any one of Claims 10 to 12, wherein there are a plurality of said grooves angularly spaced about said facing and connected in parallel with each other between said high and low pressure cavities.

14. A clutch facing as claimed in any one of Claims 10 to 13, wherein the or each groove has a rectangular cross sectional flow area with a relatively small depth and a relatively large width.

15. A clutch facing as claimed in any one of Claims 10 to 14, wherein the or each groove has a smooth sinusoidal shape with apexes located closely adjacent the edges to which the apexes are nearest.

16. A clutch facing as claimed in any one of Claims 10 to 14, wherein the or each groove has straight traversing sections and pointed return bends joining said straight traversing sections, and said return bends are located closely adjacent the edges to which the return bends are nearest.

17. A clutch facing as claimed in any one of Claims 10 to 16, wherein the or each groove has a uniform cross sectional flow area.

18. A clutch facing as claimed in any one of Claims 10 to 17, wherein there are four of said grooves equally angularly spaced about said facing and there are five of said passes.

19. A clutch facing as claimed in any one of Claims 10 to 12, wherein said groove is an uninterrupted groove.

20. A clutch facing substantially as herein described with reference to and as shown in any one of Figures 6, 7 or 12 of the accompanying drawings.