JPH07208577A - Friction ring - Google Patents

Abstract

PURPOSE: To optimize the cooling action to be obtained by the volume of cooling fluid by forming a groove with at least one throttle point over a partial length of the whole length thereof, and practically regulating the volume of the coolant capable of flowing through the groove by the throttle point. CONSTITUTION: In a friction ring to be used for wet clutch, especially a friction ring for a bridging clutch 15 for a hydrodynamic torque converter 3, the friction ring has at least one friction surface 21 provided with the outer peripheral part and the inner peripheral part, and a groove 24 for cooling is provided in the periphery of the friction surface 21, and a groove (recessed part) 26 secures the connection between the outer peripheral part and the inner peripheral part. In the friction ring with this structure, the groove 26 is formed with at least one throttle point 30 over a partial length of the whole length thereof, which regulates the volume of the coolant capable of flowing through the throttle point 30 and the groove 26.

Description

Detailed Description of the Invention

[0001]

FIELD OF THE INVENTION The present invention relates to a friction ring or lining used in a wet clutch, in particular a friction ring or lining for a lock-up clutch of a hydrodynamic torque converter. Of the type that forms a friction surface with grooves or passages for passing the cooling liquid.

[0002]

BACKGROUND OF THE INVENTION Friction rings or friction linings of this type and wet clutches with such friction rings or friction linings are described in US Pat.
It is known from US Pat. No. 543 and US Pat. No. 5,056,631. Both of these U.S. patents describe hydrodynamic torque converters with lock-up clutches in which the engaging friction surfaces are as follows: the lock-up clutch is closed. The oil flow between the chambers on either side of the ring piston, even if it is. A pump car, a turbine car, a guide car, and a lock-up clutch having a ring piston are housed in the casings of these hydrodynamic torque converters. A chamber that can be filled with oil is formed on both sides of the ring piston. In this case, the first chamber of the chambers is formed inside the friction surface of the lock-up clutch in the radial direction. At least a turbine car is provided in the second chamber. The first chamber is bounded by the ring piston and the radial wall of the casing. The oil flow serves to reduce the component heat load resulting from slippage in the lockup clutch, in particular in the area of the friction lining or the friction surface.

[0003]

SUMMARY OF THE INVENTION Therefore, an object of the present invention is to improve a conventionally known friction ring to provide a known friction ring having a groove through which a cooling liquid flows, and a wet clutch having the friction ring. Is to be optimized with respect to the cooling effect that can be obtained by the volumetric flow of liquid.
This is particularly desired to be achieved by a better heat exchange in the area of the friction surface of the wet clutch between the liquid and the component adjacent to the liquid. Furthermore, it is desirable for the high torque capacity of the wet clutch to be guaranteed by means of the inventive construction of the friction ring or the wet clutch. Furthermore, it is desirable that the friction ring or the friction disc provided with it, and thus the wet clutch, can be manufactured particularly easily and economically.

[0004]

In order to solve this problem, in the structure of the present invention, a groove for cooling is provided in the area of the friction surface of the friction ring, and the groove is formed as follows: The groove forms at least one throttling point over a partial length of its entire length, which throttling point substantially defines the volume of coolant that can flow through the groove. . The groove may in this case be such that it ensures the connection between the outer or outer circumference and the inner or inner circumference of the friction ring. The grooves are thus open radially outward and radially inward. In a particularly advantageous design, the at least one throttle point is designed in such a way that a swirl is produced in the region of the throttle point. The width and the depth of the groove area located outside such a throttled portion are preferably designed as follows, that is to say in which at least a substantially laminar flow occurs.

By designing the cooling channels or passages as in the present invention, direct cooling of the friction surfaces engaging one another can be obtained, especially in the case of continuous slip. By restricting the volumetric flow of the cooling fluid according to the invention in relation to the pressure difference between the two chambers of the lockup clutch of the torque converter, which are arranged on either side of the piston, a corresponding hydrodynamic It is possible to achieve optimum cooling over the entire operating range of the torque converter. The structure according to the present invention has the following advantages. That is to say, in the arrangement according to the invention, a relatively high pressure drop is produced in the region of the throttling due to the vortices present there, whereas the flow losses in the remaining groove region are relatively high. Extremely small based on the cross section. In the prior art mentioned at the outset, the groove is constructed in the following manner, that is to say in the groove a predominantly laminar throttling action takes place over the entire length. With such a throttling action, the volumetric flow rises linearly in proportion to the pressure, ie the pressure difference between both chambers of the lockup clutch. In the vortex-like throttling action according to the invention, the volumetric flow rises as a function of the pressure or the pressure difference. This means that the vortex-like throttling action is more advantageous. This is because in this case, a relatively large volume flow is obtained below the maximum permissible pressure over the entire pressure range that occurs in a substantially hydrodynamic torque converter.

As in the present invention, by using the narrowed portion having a relatively short length with respect to the entire length of the groove, the influence of the manufacturing error in the width and depth of the lining groove, the manufacturing error of the piston, and the like. Influence of deformation caused by driving,
And the effect of the corresponding friction surface on the flow resistance of the lining groove can be kept small or minimized.
Furthermore, since the viscosity of the oil decreases with increasing temperature, a larger volume flow and thus better cooling can be achieved with increasing temperature of the cooling oil by means of the throttle point according to the invention. The flow resistance of the throttle or lining groove for oil decreases with increasing temperature. However, this effect, which is advantageous in its own right, must be limited to the desired value of the oil volume by a corresponding design of the throttle. The pressure in the closed chamber of the lock-up clutch can no longer be maintained if the flow-through is too great.

In a particularly advantageous embodiment of the invention, the throttling action of the cooling medium at the throttling point, ie the short passage section, comprises an acute inflow and / or outflow. This guarantees a satisfactory swirl, whereby the flow resistance is only linearly related to the width or the depth of the restriction. For long passages that produce substantially laminar flow, as in the prior art of the present invention, the flow resistance is a value of the fourth power and depends on the hydraulic radius or diameter. That is, the tolerance of the groove size has an extremely large effect on the flow resistance.

The grooves or passages can be provided in the friction surface of the friction ring or friction lining as follows. That is, the friction surface forms, on the radially outer side, a substantially continuous ring area, which is interrupted only by the radially or obliquely extending throttle passage. On the radially inner side of the ring area, the groove area or the passage area has a significantly larger cross-section with respect to the cross-section of the throttle passage, so that in this area, compared to the flow resistance in the throttle passage, Only a very small flow resistance occurs. In other words, the main part of the total flow resistance of the cooling groove occurs in the area of the throttling location or throttling passage.

The throttle point is preferably arranged on the outer diameter of the lockup clutch or the friction ring. With such a structure, the crimping force or closing force of the lock-up clutch can be increased.
This is due to the following. That is to say that there is a pressure which is throttled to a significantly lower pressure level over the greatest part of the radial length of the friction surfaces which are engaged with one another, whereby the closing force can be correspondingly increased.

In a particularly advantageous configuration of the invention, the throttle points are arranged as follows, that is to say they are always in the holding range of the friction surface, and are constructed in this way. , The throttling action cannot be bypassed through the gap between the lining surface and the surface of the corresponding friction surface. This is particularly shown in FIGS. 7 and 8. The remaining lining area, which has a relatively large cross section, is not in full contact with the corresponding surface,
In the case of overflow, the total flow resistance is
It is reduced to the extent that it does not impair function. Because that
This is because the passage sections provided here only take up a very small part of the total throttle. The cooling oil channels or passages are preferably designed to be relatively small, so that the effects of lining fabrication errors and sets on the flow resistance are minimized. Desirably, the passages are configured so that there are no dead areas that will not be flushed by the cooling oil. The grooves provided in the friction surface or the friction lining can be formed by embossing or stamping.

In a preferred embodiment of the invention, the length of the throttle is of a value between 2 and 8 mm, preferably between 3 and 5 mm.

The cross-section ratio between the longitudinal extent of the groove with a large cross-section and the throttling point is between 3: 1 and 8: 1, preferably between 4: 1 and 6: 1. have. However, larger or smaller ratios are also possible, depending on the application.

[0013]

Embodiments of the present invention will now be described with reference to the drawings.

The device 1 shown in FIG. 1 has a casing 2 which houses a hydrodynamic torque converter 3. The casing 2 can be connected to a drive shaft, which can be formed by a driven shaft of an internal combustion engine, for example a crankshaft. The non-rotatable connection between the drive shaft and the casing 2 can be made via a drive plate, which is radially inwardly connectable with the drive shaft and non-rotatable. It can be connected non-rotatably to the casing 2 on the outside. Such a drive plate is disclosed in, for example, Japanese Patent Laid-Open No. 58-30
It is known based on Japanese Patent No. 532.

The casing 2 is formed by a casing shell 4 adjacent to the drive shaft or adjacent to the internal combustion engine, and another casing shell 5 fixed to the casing shell. Both casing shells 4, 5 are rigidly sealingly connected to each other via a welded joint 6 on the radially outer side. In the embodiment shown, the casing shell 5 is used directly to form the outer shell of the pump wheel 7. For this purpose, the blade plate 8 is fixed to the casing shell 5 in a manner known per se. The casing shell 5 is fitted axially in a sleeve-like area 4a outside the casing shell 4. A turbine wheel 10 is provided between the pump wheel 7 and the radial wall 9 of the casing shell 4 when viewed in the axial direction, which turbine wheel 10 is rigid or non-rotatable with the driven hub 11. The driven hub 11 is coupled and is non-rotatably coupled to the transmission input shaft via internal teeth. A guide wheel 12 is provided between the radially inner area of the pump vehicle and the radially inner area of the turbine vehicle when viewed in the axial direction. The casing shell 5 has a sleeve-like hub 13 on the inside in the radial direction, which hub 13 can be rotatably and sealingly supported in the casing of the transmission. A lock-up clutch 15 is further provided in the inner chamber 14 formed by both casing shells 4 and 5, which lock-up clutch 15 is operatively arranged in parallel with the torque converter 3. It is arranged. The lockup clutch 15 enables torque coupling between the driven hub 11 and the drive-side casing shell 4. In operation, a rotationally elastic damper 16 is connected in series with the lockup clutch 15. In the illustrated embodiment, the damper 16 is a ring-shaped piston 17 of the lockup clutch 15 and a driven hub. It is provided between 11 and. The rotationally elastic damper 16 has a power store in the form of a coil spring in a manner known per se. Viewed in the axial direction, the ring-shaped piston 17 provided between the radially extending wall 9 and the turbine wheel 10 is slid in the axial direction while being restricted along the driven hub 11 on the radially inner side. Possibly supported on the driven hub 11. The ring-shaped piston 17 has an inner chamber 14
Is divided into a first chamber 18 and a second chamber 20. In this case, the first chamber 18 has a ring shape when viewed in the axial direction inside the friction engagement range 19 of the lockup clutch 15. Formed between the piston 17 and the radial casing wall 9, the second chamber 20 accommodates therein the pump wheel 7, turbine wheel 10 and guide wheel 12, in particular.

The casing shell 4 forms a friction surface 21 in a radially outer ring-shaped area, which friction surface 21 and a friction lining 22 held by a ring-shaped area 23 of the piston 17. Can be engaged.

In a new concept, for example for motor vehicle powertrains, the lock-up clutch is operated with slip over at least a large part of the operating range of the fluid converter, in which case in the friction engagement range 19 during the slip phase. In the form of heat, a loss output occurs, which is extremely high under certain operating conditions and can reach several kilowatts. Such an operating state exists, for example, when traveling on a slope accompanied by a towed vehicle or when the converter clutch is switched from a non-locked-up state to a substantially locked-up state. Such a concept of operating a converter lockup clutch with slippage is described, for example, in German Patent Application No. 432818.
No. 2.6 proposed.

In order to avoid unacceptably high temperatures in the frictional engagement area 19, and thus at least to prevent the destruction of the friction lining surface and the destruction of some of the oil in the inner chamber 14, in the embodiment shown. Means in the form of oil grooves or passages 24 provided in the friction lining 32 are provided, through which the second lock-up clutch 15 can also be substantially closed. There is always an oil flow between the chamber 20 and the first chamber 18. In this case, the oil flow is guided via the friction surface 22 a of the friction lining 22 and the friction surface 21. The oil passage 24 is optimally configured in its shape to ensure good heat exchange between the friction-engaging member in the area 19 and the flowing oil. The advantageous shaping of the passage 24 will be described in detail with reference to FIGS.

The end of the passage 24 located further outward in the radial direction is connected to the chamber 20 and the passage 24
The end located further inside in the radial direction of the
Connected with. Locked up clutch 1 closed
At 5, the cooling oil flow enters the chamber 18 via the passage 24 and flows radially in this chamber 18 towards the axis of rotation 25.

This cooling oil flow is then driven by the driven hub 1.
In one area, it can be led out, for example, via a hollow shaft or a passage provided therefor, that is to say preferably in the oil cooler. The oil is returned from the oil cooler to the oil sump, and then from the oil sump to the hydraulic type adjusting circuit or control circuit.

FIG. 2 shows a circular ring-shaped friction lining 2
2 is partially shown, this friction lining 22
Can be used in the converter lockup clutch shown in FIG. The friction lining 22 is
It has grooves or recesses 26 distributed over the entire circumference, and these grooves or recesses 26 are provided in both chambers 18,
A connection passage 24 between the 20 is formed.

The friction lining 22 has an outer peripheral portion or outer diameter portion 27 and an inner peripheral portion or inner diameter portion 28. A short section 29 of the passage 24 connecting the outer diameter portion 27 and the inner diameter portion 28 forms a throttling point 30. This section 29 of the passage 24 is in this case radially oriented and, on the radially inner side, transitions into a circumferentially extending radially outer passage section 31, which passage section 31.
Via a hairpin-shaped deflecting section 32 to a passage section 33 located further inward in the radial direction and likewise extending in the circumferential direction. The passage section 33 is connected to a radially inwardly open outflow region 34 for the cooling medium guided via the passage 26. The cross-sections of the passage areas or passage sections 31, 32, 33, 34 which connect to the throttling point 30 are as follows in relation to the cross-section of the throttling point 30, that is to say of the device of FIG. It is arranged such that even if the pressure difference between the two chambers 18, 20 is maximal, there is essentially a laminare Stroemung. In the region of the throttle portion 30, a vortex flow is substantially constantly generated during the operation of the device shown in FIG. 1 and when the lockup clutch 15 is frictionally engaged. That is, the passage 24 is constructed as follows. That is, the volume of the cooling liquid flowing through the passage 24 is not defined by the flow resistance created over the entire length of the passage as in the prior art, but rather mainly by the throttling point 3.
It is defined by the resistance in the range of 0. As can be further seen from FIG. 2, the circumferentially extending passage sections 31, 33 have sections 35, 36 which, in connection with the throttle section 30, extend in different rotational directions. ing. The sections 35, 36 are in this case arranged symmetrically with respect to the throttling point 30 when viewed in the circumferential direction.

As shown in FIGS. 2-4, the length 29 of the throttled portion 30 is a very small portion of the total length of the passage 24. The outer diameter portion 27 has a size of 180 to 260 m
Due to the general structural dimensions of the wet clutch (Nasslaufkupplung) or the lockup clutch 15, which is between m, the length 29 of the throttling point 30 is between 2 and 8 depending on the application.
mm, preferably between 3 and 5 mm.

The flow-through cross section of the throttling point 30 has groove sections 31, 32, 33, which are connected to the outlet side of the corresponding throttling point 30.
It is several times smaller than the flow-through cross section of 34. The dimensional ratio in this case is between 1: 3 and 1:10. For many applications, however, ratios between 1: 4 and 1: 6 are sufficient.

Due to the significantly larger flow cross-sections of the passage sections 31, 32, 33, 34, it is ensured that laminar flow occurs in these sections virtually always or exclusively.

In order to obtain an optimum swirl in the region of the throttling point 30 formed by the passage-like short recess,
In an advantageous configuration of the invention, at least the cross-section in the inflow range of the throttling point 30 has an acute angle. Figure 2
In the illustrated embodiment shown in FIG. 3, the throttled portion 30 transitions on the outflow side into a corresponding groove region 31 via a divergent widening formed by a round chamfer. However, in a further advantageous configuration, an acute cross-section transition is provided between the passage section 31 and the throttle point 30.

The area required by the groove or passage 24 is 30 to 65%, preferably 40 to 55%, of the area between the outer diameter portion 27 and the inner diameter portion 28. In the embodiment shown in FIG. 2, this ratio is approximately 50%.

The throttling point 30 is preferably constructed as follows. Thus, in this case, the throttling point 30 ensures that at least approximately 60 to 85%, preferably 70 to 80%, of the maximum pressure difference between the two chambers 18, 20 is eliminated. This means that after the throttling point 30 or immediately after the throttling point 30, the pressure in the passage section 31 is approximately 15-40% or 20-30% greater than the pressure in the chamber 18. Based on the mode of action of the throttling point 30, as shown in FIG.
They are preferably arranged radially outside the friction ring 22, that is to say in the high-pressure range. This is because the friction surface 2
This is because the pressure generated in the range of 1 and 22a and acting against the closing pressure of the lockup clutch 15 can be kept small. This allows the transferable moment from the lockup clutch 15 due to the pressure difference between the two chambers 18, 20 to be compared with known lockup clutches with cooling passages and a corresponding amount of cooling liquid. Can be increased. In this case, however, at least some throttle points 30 are displaced radially inward so that both chambers 18,
The torque capacity of the lock-up clutch 15 can also be reduced due to the differential pressure provided during 20.

FIG. 5 shows the influence of the radial arrangement of the throttle points 30 on the torque that can be transmitted. FIG. 5 shows on the left side a partial area between the casing shell 4 and the piston 17, together with a friction lining 22 fixed to the piston 17. On the right side of FIG. 5, an ideal pressure pattern (Druckprofil) that can be realized with respect to the radial length range of the friction lining 22 and the arrangement form of the throttle portions is shown. Relatively high pressure p1 in chamber 20 and relatively low pressure p2 in chamber 18
In view of the radial length of the friction lining 22, the friction lining 22 has a radial extension, as in the case shown in FIG. In the area between the friction surface 22a and the friction surface 21, a pressure distribution extending in the passage 24 is possible, as shown by the chain line 37. As is apparent from the alternate long and short dash line 37, approximately 80% of the pressure difference between p1 and p2 in the range of the throttled portion 30.
Is extinguished. The pressure difference between the pressure Pa in the vicinity of the outlet side of the throttling point 30 and the pressure p2 in the chamber 18 is therefore relatively small. When the throttle portion 30 is arranged radially inward, that is, when the throttle portion 30 is arranged in the range of the outflow section 34 as shown in FIG. 2, the engagement range 19 is indicated by the broken line 38. Such a pressure distribution is produced. As can be seen from the dash-dotted line 37 and the dashed line 38, the lock-up clutch 15 is owing to the pressure difference provided between the two chambers 18, 20.
For the moment that can be transmitted by
It can be influenced by placing 0 at different diameters. By arranging the throttling points 30 radially outwards, the differential pressure between the two chambers 18, 20 necessary for transmitting the defined moment is equal to the cooling oil flow between the two chambers 18, 20. It can be reduced compared to the conventional lock-up clutch equipped with. Depending on the number and shape of the throttling points 30, the flow-through width of such throttling points 30 is a value between 0.4 and 2.5 mm,
It can advantageously have a value between 0.5 and 1.5 mm. The depth of the groove 26 can have a value between 0.2 and 1 mmno, preferably a value between 0.3 and 0.7 mm.
The depth of the groove 26 can be substantially equal over its entire length. However, it is also possible for the grooves 26 to have different depths. It is advantageous if a relatively large depth is provided, in particular in the region of the throttle point 30 and possibly in the transition range between the throttle point 30 and the remaining groove segment 31. This is indicated by reference numeral 30 in FIG.
It is indicated by the dash-dotted line labeled a. That is, in an advantageous configuration of the invention, the throttling points are designed to be somewhat deeper than the other groove areas and, in compensation, to have a somewhat smaller width in order to obtain the desired cross-section of the throttling points 30. Has been done. With such a configuration, it is possible to reliably reduce the change in the throttling action of the throttling portion 30, which is associated with the wear of the friction lining 30, that is, the wear of the friction lining 30 that causes a reduction in the cross section of the throttling portion 30. it can.

Thus, according to the invention, the volumetric flow of the cooling fluid in the wet clutch is regulated by means of at least one throttle 30, in which case the remaining lining surface has a corresponding throttle in the flow direction. Relatively long passages can be formed behind the points 30 and these passages are
Ensures as little flow resistance as possible and a large heat exchange surface.

FIG. 6 shows the pressure difference p1-p2 (Δp) between the two chambers 20 and 18 on the horizontal axis. The vertical axis shows the volume flow that occurs in relation to the pressure difference.

If the volumetric flow is laminarly throttled over the length of the groove formed in the friction lining, there is a substantially linear or linear relationship between the pressure difference in the groove and the volumetric flow. Existing. This relationship is shown in Figure 6 by a straight solid line. The pressure difference in a groove is the difference between the pressure on the inlet side and the pressure on the outlet side of the corresponding groove. Such a laminar flow-like throttling action is achieved by the prior art of the type mentioned at the outset, namely US Pat. No. 4,696,543 and US Pat.
This occurs in the groove construction described in the '631 specification. In this prior art, the laminar flow throttling is approximately 70% of the total throttling performed in the passage.

The dashed line in FIG. 6 shows the volume flow which can be obtained by the vortex-like throttling action according to the invention. The volume flow course associated with the pressure difference between the two chambers 20, 18 is essentially the root function (Wurz
elfuktion). The profile shown in broken lines can be obtained by the inventive design of the groove, in particular by the design shown in FIG. As can be seen from FIG. 6, a larger volumetric flow can be used than in the laminar flow throttling action, especially when the pressure difference is small in the vortex flow throttling action. This is particularly advantageous. This is because in order to ensure the best possible cooling, it is desirable to be able to use as large a volumetric flow as possible in the locked-up state even when the pressure difference between both chambers 20, 18 is small. Is.

The configuration of the groove corresponding to the two characteristic lines shown in FIG. 6 is as follows: that the groove configuration ensures an equal volume flow for a predefined Δp max. , Designed. This Δp max is a value between 7 and 10 bar in a general-purpose hydrodynamic torque converter with a lockup clutch. However, Δp max may also lie above or below this bandwidth.

The arrangement according to the invention of the grooves or passages provided for the cooling oil flow further reduces the temperature-dependent variation of the flow-through through these passages. This is because, in the present invention, the main throttling action, that is, the pressure loss, is performed within a relatively short throttling area. The grooves in the region of the throttling point influence the flow-through or volumetric flow only in a straight line, which ensures that they are less dependent on geometrical tolerances. In the above-mentioned conventional groove structure,
Most throttling takes place in laminar flow, that is to say over the entire length of the groove. In such a throttling action, the groove height is a value of the fourth power, which affects the flow amount or the volume flow. This increases the dependence on the geometrical tolerances of the lining or groove. Furthermore, in the prior art, the volume flow has a large dependency on the viscosity or the temperature of the cooling medium due to the laminar flow-like throttling action.

In order for the groove according to the invention to always ensure its throttling action, the contact of the friction lining 22 on the corresponding friction surface 21 is necessary in the region of the throttling point.
In this case, at least the following is ensured, that is, no gap exists in the range of the throttled portion under any operating condition, or if any, the gap is 0.03 mm or less. It is preferably 0.01 mm or less. Such gaps can occur due to poor parallelism between the friction surfaces engaging each other.

In order to ensure that the throttling point 30 fulfills its function in all operating conditions in which the friction surfaces are engaged, the friction lining 22, as shown in FIG. Component or ring piston 17
It is advantageous to be held by.

FIGS. 7 and 8 show a magnified partial area of the casing shell 4 and the piston 17 and of the friction lining 22 fixed to the piston 17. FIG. 7 shows the configuration in which the piston 17 occupies a substantially unloaded or relaxed state. This piston shape is designed for both chambers 18, 20
At substantially equal pressures, or if there is only a relatively small pressure difference. In the relaxed state of the piston 17, the outer area 17a with the friction lining 22 is as follows. That is, in this case, the friction surface 22a of the friction lining 22 and the friction surface 21 of the casing 4 form a wedge-shaped void 39 therebetween, and the void expands radially inward to form an angle φ. Have This angle φ is for example 0.5 to 3 °, preferably 1 °.

FIG. 8 shows a case where a predetermined overpressure is generated in the chamber 20 with respect to the chamber 18,
The position of the piston 17 is shown. This overpressure is 4 ~
Values of between 8 bar, in which case the piston 17 must be correspondingly spring-elastic, depending on the desired maximum overpressure.

As can be seen in both FIGS. 7 and 8, when the pressure between the two chambers 18, 20 is small or the differential pressure is small, the friction lining 22 is simply provided with a throttling point 30. It is in frictional contact with the friction surface 21 via a radially outer ring-shaped friction surface section 40 which is provided. This allows the following: both chambers 1
If the differential pressure between 8 and 20 is already small, or if the overpressure in the chamber 20 is small (eg 1 bar), the throttling point 30 takes over its function, but is guaranteed. As the overpressure on chamber 18 in chamber 20 increases, piston 17 deforms from the shape shown in FIG. 7 to the shape shown in FIG. By this, both friction surfaces 2
The contact range between the two frictional surfaces 21 and 22a gradually increases, that is, the angle φ between the two friction surfaces 21 and 22a decreases. However, the throttling point 30 also ensures a satisfactory control of the coolant volume.

The friction lining or friction ring 22 shown in FIG. 2 is integrally constructed. However, the friction lining or the friction ring may also consist of sector-shaped individual lining parts which are joined to one another in the circumferential direction.

The friction linings 122,
222 and 322 are partially shown, these friction linings being provided with grooves or passages according to the invention.

The friction linings shown in FIGS. 9 to 11 all have throttle points 130, 230, 330 distributed over the entire circumference of the friction lining. These throttling points 130, 230, 330 exclusively define the volumetric flow that can flow through the passages 124, 224, 324. Throttling point 13
The passage sections connected to 0, 230, 330 have a significantly larger flow cross-section than the throttling points 130, 230, 330, so that predominantly laminar flow is present in these passage sections. Passages 124, 224, 324
The flow velocities in these sections of
It is significantly smaller than the flow velocity at 30, 230 and 330. This achieves optimum heat transfer between the flowing cooling medium or cooling oil and the adjacent components.

In the configuration shown in FIG. 9, the friction lining 1
The reference numeral 22 has a ring-shaped groove 131 connected to the narrowed portion 130, and this groove 131 is connected to a large number of radial grooves 132 extending inward. The friction surface of the friction lining 122 has individual grooves 132.
Is formed by a ridge 132a existing between the ridges and a ring-shaped ridge 122a existing in the edge region of the friction ring 122. The ring-shaped ridge 122a is formed by:
It is divided into individual sector-shaped sections by the narrowed portion 130.

The friction lining 222 shown in FIG.
Is a plurality of ring-shaped recessed portions 231, 231a, 231
b, and these recessed portions 231, 231a, 2
31b are connected to each other by radially extending groove areas 232, 232a. The ring-shaped groove area 231b on the inner side in the radial direction is open toward the inner side in the radial direction via the groove area 232b in the radial direction. The radial groove areas 232, 232a, 232b are circumferentially offset from each other such that the oil flowing through the passage 224 is redirected many times.

In the embodiment shown in FIG. 11, the passage 324
Has a meandering shape in the circumferential direction at the connection to the throttle portion 330. As a result, based on the area and the length of the meandering range of the passage 324, the component adjacent to the cooling oil or the adjacent friction is formed. There is good heat exchange between the faces.

According to another embodiment of the present invention, it is also possible to provide cooling grooves constructed as in the present invention in the area of the friction surface 21 of the casing 4 instead of being mounted on the friction lining 22. And such a cooling groove,
It can be formed by embossing a sheet metal material. On the radially outer side and the radially inner side, the embossed passages must be constructed as follows: the passages open towards the chambers 18,20. Furthermore, the friction lining 22 is instead of being retained by the piston 17 and instead of the casing 4
It may be fixed to. Furthermore friction lining 2
The 2 may be held by an intermediate plate, as in some embodiments in the prior art. The cooling passage according to the invention may furthermore be embossed directly in the material forming the piston 17, in which case the friction lining 22 is held by the casing 4 or an intermediate plate.

The grooves or passages formed in the friction lining or the friction ring are
That is, it can be formed prior to fixing the friction lining to a retaining component such as a ring piston or plate. However, it is also possible for the grooves or channels to be formed in the friction lining during the fixing of the friction lining to the holding component, for example by gluing, or to the friction lining after such fixing. Is. For example, the friction lining 22 shown in FIG.
The passage 24 can first be fixed to the ring piston 17 and the passage 24 can be embossed into the friction ring 22 during or after this fixing. This is done by means of a pressing tool with a corresponding molding.

[Brief description of drawings]

FIG. 1 is a cross-sectional view showing an apparatus including a wet clutch having a friction ring according to the present invention.

FIG. 2 is a partial view of a friction lining constructed as in the present invention.

FIG. 3 is an enlarged cross-sectional view taken along line III-III in FIG.

FIG. 4 is an enlarged cross-sectional view taken along line IV-IV of FIG.

FIG. 5 is a diagram showing a pressure distribution in a radial direction that can be obtained when a narrowed portion is arranged in the range or groove of the friction surface as in the present invention.

FIG. 6 is a diagram showing the function and effect of the groove improved by the structure of the present invention.

FIG. 7 shows an arrangement for the friction lining, with the piston not loaded.

FIG. 8 is a diagram corresponding to FIG. 7, showing a case where overpressure occurs in one chamber.

FIG. 9 is a diagram showing an embodiment of a cooling passage or a cooling groove.

FIG. 10 is a diagram showing another embodiment of a cooling passage or a cooling groove.

FIG. 11 is a view showing still another embodiment of the cooling passage or the cooling groove.

[Explanation of symbols]

1 device, 2 casings, 3 torque converter, 4,5 casing shells, 6 welded joints,
7 pump cars, 8 blade plates, 9 walls, 1
0 turbine wheel, 11 driven hub, 12 guide wheel, 13 hub, 14 inner chamber, 15 lock-up clutch, 16 damper, 17 piston, 1
8, 20 chambers, 19 friction engagement range, 21 friction surface, 22, 122, 222 friction lining, 23
Ring-shaped range, 24,124,224,324
Oil groove (oil passage), 25 rotation axis, 26
Recessed portion, 27 outer diameter portion, 28 inner diameter portion, 30, 13
0,230,330 throttling points, 31 passage sections,
32 turning portion, 33 passage section, 34 outflow range,
35, 36 divisions, 122a, 132a ridges

Claims (18)

[Claims] 1. A friction ring for use in a wet clutch, in particular for a lock-up clutch of a hydrodynamic torque converter, the friction ring comprising:
In this case, the friction ring has at least one friction surface having an outer peripheral portion and an inner peripheral portion, and a groove for cooling is provided in the range of the friction surface, and the groove has an outer peripheral portion and an inner peripheral portion. In the type that ensures a connection with the circumference, the groove forms at least one throttling point over a partial length of its entire length, which throttling point flows through the groove. Friction ring characterized by defining the volume of possible cooling liquid. 2. The friction ring according to claim 1, wherein the throttling points are designed for vortex flow and the remaining groove areas are designed for substantially laminar flow. 3. Friction ring according to claim 1 or 2, wherein the length of the constriction has a value of 2 to 8 mm, preferably 3 to 5 mm. 4. The cross-section ratio between the longitudinal extent of the groove with a relatively large cross-section and the throttling point is 3: 1 to 8 :.
Friction ring according to any one of claims 1 to 3, having a value between 1 and preferably between 4: 1 and 6: 1. 5. The friction ring according to claim 1, wherein the narrowed portion is formed by a short passage-shaped recessed portion and is provided with an acute inflow portion and / or an outflow portion. . 6. The friction ring according to claim 1, wherein at least one throttle portion is provided on an outer peripheral portion of the friction ring. 7. The friction ring has throttling points extending radially from the outer peripheral portion and distributed over the entire circumference, the throttling points being provided in groove sections extending in the circumferential direction. Friction ring according to any one of claims 1 to 6, which is transitioning and in which the groove section is connected radially inward with an outflow section which opens towards the inner edge of the friction ring. 8. The throttle point is transitioned into a circumferentially extending radially outer groove section, which groove section extends through the radially extending groove section in the circumferential direction. 8. The friction ring of claim 7, further comprising: an inner groove section open to an outflow section. 9. The friction ring according to claim 7, wherein the circumferentially extending groove sections are arranged symmetrically with respect to the assigned throttling point when viewed in the circumferential direction. 10. A friction ring according to any one of claims 7 to 9, wherein the outflow cross section is located opposite the throttled part when viewed in the radial direction. 11. The friction ring is provided with a plurality of throttling points distributed over the entire circumference thereof, the throttling points transitioning into a groove section extending in a zigzag or meandering manner in the circumferential direction. The friction ring according to any one of 1 to 6. 12. A friction ring according to claim 1, wherein the groove has at least two deflections. 13. The area occupied by the groove is 30 to 60%, preferably 40 to 50%, relative to the area between the outer circumference and the inner circumference of the friction ring. The friction ring according to any one of 1 to 12. 14. The friction ring according to claim 1, wherein the groove is formed in the friction ring by embossing or cutting. 15. The friction ring is a component of a lock-up clutch of a hydrodynamic torque converter, in which case the torque converter has a casing in which the pump car, the turbine car and the guides are guided. It houses a car and a lockup clutch,
The lockup clutch has a ring piston,
A chamber that can be filled with oil is provided on each side of the ring piston, the ring piston having at least one friction surface, the friction surface being frictionally engageable with a corresponding friction surface, In this case, a first chamber of the chambers is formed radially inside the friction surface between the ring piston and the component having the corresponding friction surface, and at least one of the ring surfaces comprises To 13 and in this case, when the friction surfaces are in axial contact, a pressure difference between the two chambers is generated via a groove provided in the friction ring. A friction ring characterized by an oil flow based on it. 16. In the region of the restriction, the pressure difference between the two chambers is approximately 60-80%, preferably 70-.
16. The friction ring of claim 15, which is 80% extinguished. 17. A friction ring according to claim 15 or 16, wherein the throttling point is adjacent to a chamber having a relatively high pressure when the lockup clutch is closed. 18. The friction ring according to claim 15, wherein the groove is formed in the friction ring by embossing or cutting.