Friction clutch

A friction clutch for use in motor vehicles wherein the pressure plate is automatically shifted toward the counterpressure plate for the friction linings of the clutch disc in dependency on the extent of wear upon the counterpressure plate, pressure plate and particularly the friction linings. The pressure plate is non-rotatably but axially movably coupled to the housing of the friction clutch and is biased toward the friction linings by resilient means including a diaphragm spring or a set of coil springs. The adjusting unit which changes the position of the pressure plate as a function of the extent of wear mainly upon the friction linings can employ cooperating wedges including a set of wedges on the housing and a set of wedges on the resilient means.

BACKGROUND OF THE INVENTION 
The invention relates to improvements in friction clutches, especially for 
use in motor vehicles. More particularly, the invention relates to 
improvements in friction clutches of the type wherein a pressure plate is 
non-rotatably but axially movably connected to a housing or cover and is 
biased by one or more prestressed resilient devices (e.g., in the form of 
diaphragm springs) to urge a clutch plate or clutch disc against a rotary 
counterpressure plate (e.g., a flywheel) which is driven by the output 
element of an internal combustion engine in a motor vehicle. The clutch 
disc can serve to transmit torque to the input element of a variable-speed 
transmission in the power train between the engine and the wheels of a 
motor vehicle. 
Friction clutches of the above outlined character are disclosed, for 
example, in published German patent application Serial No. 24 60 963, in 
German Pat. No. 24 41 141, in German Pat. No. 898 531 and in German 
Auslegeschrift No. 1 267 916. 
OBJECTS OF THE INVENTION 
An object of the invention is to prolong the useful life of friction 
clutches, particularly of friction clutches for use in motor vehicles. 
Another object of the invention is to provide a friction clutch whose mode 
of operation is less affected by wear upon its parts than in heretofore 
known friction clutches. 
A further object of the invention is to provide a friction clutch which can 
be disengaged in response to the application of a relatively small force. 
An additional object of the invention is to provide a friction clutch which 
is constructed and assembled in such a way that the magnitude of the 
disengaging force need not change, or changes negligibly, irrespective of 
the extent of wear upon certain component parts of the friction clutch. 
Still another object of the invention is to provide a simple, compact and 
inexpensive friction clutch which embodies the aforediscussed features. 
A further object of the invention is to provide a novel and improved method 
of compensating for wear upon the clutch plate or clutch disc in a 
friction clutch of the above outlined character. 
Another object of the invention is to provide a novel and improved method 
of compensating for wear upon the clutch disc and/or certain other 
component parts of a friction clutch for use in motor vehicles and the 
like. 
An additional object of the invention is to provide the above outlined 
friction clutch with a novel and improved system which can automatically 
compensate for wear upon the friction linings forming part of the clutch 
disc. 
Still another object of the invention is to provide a fiction clutch which 
exhibits the above outlined advantages and which can be assembled at a 
reasonable cost in automatic machines. 
A further object of the invention is to provide a novel and improved 
aggregate which employs the above outlined friction clutch as well as one 
or more additional components, such as the part or parts which transmit 
torque to the input element and/or receive torque from the output element 
or elements of the friction clutch. 
Another object of the invention is to provide a motor vehicle which 
embodies the above outlined friction clutch. 
An additional object of the invention is to provide a preassembled modular 
engageable and disengageable aggregate which embodies the above outlined 
friction clutch. 
Still another object of the invention is to provide a driving unit which 
can be used in a motor vehicle and embodies the above outlined friction 
clutch. 
A further object of the invention is to provide a novel and improved 
diaphragm spring for use in the above outlined friction clutch. 
Another object of the invention is to provide a novel and improved clutch 
plate or clutch disc for use in the above outlined friction clutch. 
An additional object of the invention is to provide a novel and improved 
method of installing a diaphragm spring in the housing or cover of the 
above outlined friction clutch. 
Still another object of the invention is to provide the above outlined 
friction clutch with novel and improved means for tiltably mounting the 
diaphragm spring for the pressure plate in the housing of the friction 
clutch. 
A further object of the invention is to provide a novel and improved 
connection between the input element of the above outlined friction clutch 
and the output element of an internal combustion engine. 
Another object of the invention is to provide novel and improved means for 
biasing the pressure plate in a friction clutch for use in motor vehicles. 
An additional object of the invention is to provide friction clutch whose 
operation is not affected by wear (even extensive wear) upon friction 
linings, pressure plate and/or other parts which are subject to wear when 
the friction clutch is in use. 
Still another object of the invention is to provide a preassembled friction 
clutch which can be utilized in existing motor vehicles and/or for other 
purposes as a superior substitute for existing friction clutches. 
A further object of the invention is to provide the above outlined friction 
clutch with novel and improved means for opposing the bias of the 
diaphragm spring for the pressure plate. 
Another object of the invention is to provide a novel and improved 
combination of friction clutch and engine for use in a motor vehicle. 
SUMMARY OF THE INVENTION 
One feature of the present invention resides in the provision of an 
engageable and disengageable torque transmitting friction clutch which can 
be utilized with advantage in vehicles, particularly motor vehicles. The 
improved friction clutch comprises a housing or cover which is rotatable 
about a predetermined axis, a pressure plate, means (e.g., a group of leaf 
springs) for non-rotatably connecting the pressure plate to the housing 
with limited freedom of movement in the direction of the predetermined 
axis, a rotary counterpressure plate (e.g., a flywheel which is driven by 
the output element of an engine in a motor vehicle) adjacent the pressure 
plate, a torque transmitting clutch disc between the two plates, and at 
least one resilient device reacting against the housing to bias the 
pressure plate toward the counterpressure plate in order to clamp the 
clutch disc against the counterpressure plate and to thus rotate the 
clutch disc about the predetermined axis. The clutch disc has friction 
linings which are engageable by and disengageable from at least one of the 
two plates and are subject to wear as a result of repeated engagement with 
and disengagement from the at least one plate, and the friction clutch 
further comprises an adjusting unit including means for compensating for 
wear upon the friction linings to thereby maintain the bias of the at 
least one resilient device upon the pressure plate at a substantially 
constant value, means for engaging and disengaging the friction clutch 
including actuating means movable along a predetermined path to engage and 
disengage the friction clutch, and means for varying the torque 
transmitted by the friction clutch and/or by the clutch disc including 
means for gradually reducing the transmitted torque at least during a 
portion of movement of the actuating means along the predetermined path to 
disengage the friction clutch. 
The pressure plate comprises a portion which is engaged and biased by the 
at least one resilient device, and the disengagement of the friction 
clutch can involve axial movement of the pressure plate away from the 
counterpressure plate against the bias of the at least one resilient 
device. The friction clutch can further comprise means for gradually 
reducing the torque which is transmittable by the friction clutch at least 
during a portion of axial movement of the pressure plate. 
The friction clutch can also comprise means for securing the housing to the 
counterpressure plate to thus establish a power train between the 
actuating means and the securing means. The torque varying means can be 
disposed in the power train. 
The pressure plate has a friction surface which is engageable with the 
friction linings to establish a power train between the actuating means 
and the clutch disc, and the torque varying means can be disposed in such 
power train. 
The friction linings can include a first and a second set of friction 
linings, and the torque varying means can be disposed axially between the 
two sets of friction linings. 
The torque varying means can include means for axially yieldably locating 
at least one of the two plates and the friction linings relative to the 
others of the two plates and the friction linings, and the torque varying 
means can be acted upon by a variable force which decreases to a minimal 
value in response to disengagement of the friction clutch and gradually 
increases to a maximum value at least during a portion of movement of the 
actuating means along the predetermined path to engage the friction 
clutch. 
The means for varying the torque which is transmittable by the friction 
clutch can include means for reducing the transmitted torque during 
approximately 40-70 percent of movement of the actuating means along the 
predetermined path in a direction to disengage the friction clutch and for 
gradually increasing the torque which is transmittable by the friction 
clutch during approximately 40-70 percent of movement of the actuating 
means along the predetermined path in a direction to engage the friction 
clutch. 
The at least one resilient device (such as a diaphragm spring) can have a 
degressive force-to-displacement ratio at least during a portion of 
movement of the actuating means along the predetermined path in a 
direction to disengage the friction clutch. 
As mentioned above, the at least one resilient device can comprise a 
diaphragm spring which bears against the pressure plate. The friction 
clutch preferably further comprises a seat which tiltably mounts the 
diaphragm spring in the housing. The diaphragm spring can comprise an 
annular portion and the actuating means can comprise prongs or tongues 
which extend from the annular portion of the diaphragm spring. Such prongs 
can be of one piece with the annular portion of the diaphragm spring. The 
seat can comprise two portions (e.g., in the form of wire rings) which are 
disposed at opposite sides of the diaphragm spring. The latter can have a 
substantially sinusoidal force-to-displacement characteristic curve 
including a maximum, a minimum, a degressive portion between the minimum 
and the maximum, an operating point at the degressive portion in engaged 
condition of the friction clutch, and a ratio of forces from approximately 
1:0.4 to 1:0.7 between the maximum and the minimum. 
If the friction clutch is used in a motor vehicle, the means for engaging 
and disengaging the friction clutch can further comprise means for moving 
the actuating means and such moving means can include or constitute a 
pedal which is similar or analogous to the gas pedal of the motor vehicle. 
Another feature of the invention resides in the provision of a motor 
vehicle having a gas pedal, an engageable and disengageable friction 
clutch, means for engaging and disengaging the friction clutch including 
actuating means movable along a predetermined path to disengage the 
clutch, and means for moving the actuating means including a second pedal 
which is similar or analogous to the gas pedal of the motor vehicle. 
An additional feature of the invention resides in the provision of a 
preassembled engageable and disengageable clutch assembly or aggregate 
which comprises a housing, a pressure plate, a counterpressure plate which 
is rotatable about a predetermined axis, means for non-rotatably 
connecting the pressure plate to the counterpressure plate with limited 
freedom of movement in the direction of the predetermined axis, a torque 
transmitting clutch disc between the two plates, at least one resilient 
device which reacts against the housing to bias the pressure plate toward 
the counterpressure plate and to thereby clamp the clutch disc between the 
two plates, friction linings which form part of the clutch disc and are 
engageable by and disengageable from at least one of the two plates so 
that they are subject to wear as a result of repeated engagement with and 
disengagement from the at least one plate, an adjusting unit including 
means for compensating at least for wear upon the friction linings to 
thereby maintain the bias of the at least one resilient device upon the 
pressure plate at a substantially constant value, means for engaging and 
disengaging the clutch aggregate or assembly including actuating means 
movable along a predetermined path to disengage the clutch aggregate or 
assembly, and means for gradually reducing the torque which is transmitted 
by the clutch disc during a portion of movement of the actuating means to 
disengage the clutch assembly or aggregate. The torque reducing means can 
include at least one resilient element which is in series with the at 
least one resilient device. 
Still another feature of the invention resides in the provision of a clutch 
aggregate or assembly which comprises a twin-mass flywheel including a 
first rotary mass connectable to an output shaft of a combustion engine 
and a second mass rotatable relative to the first mass, an oscillation 
damper having means for opposing rotation of the two masses relative to 
each other, and a torque transmitting friction clutch including a 
counterpressure plate forming part of the second mass, a pressure plate, 
means for non-rotatably connecting the pressure plate to the 
counterpressure plate with limited freedom of axial movement, a torque 
transmitting clutch disc between the two plates, at least one resilient 
device acting upon the pressure plate to bias the clutch disc against the 
counterpressure plate, friction linings forming part of the clutch disc 
and being engageable with and disengageable from at least one of the two 
plates and being subject to wear as a result of repeated engagement with 
and disengagement from the at least one plate, an adjusting unit including 
means for compensating at least for wear upon the friction linings to 
thereby maintain the bias of the at least one resilient device upon the 
pressure plate at a substantially constant value, means for engaging and 
disengaging the friction clutch including actuating means movable along a 
predetermined path to disengage the friction clutch, and means for 
gradually reducing the torque which can be transmitted by the friction 
clutch and/or by its clutch disc during a portion of movement of the 
actuating means in a direction to disengage the friction clutch. 
The friction clutch of the just outlined clutch aggregate or assembly can 
further comprise a housing and means for securing the housing to the 
second mass so that the housing is separable from the second mass only in 
response to at least partial destruction or deformation of one of the two 
parts including the housing and the second mass. 
The clutch disc of the aforementioned clutch aggregate or assembly can be 
provided with at least one substantially annular friction surface and the 
damper can be located radially outwardly of the friction surface. 
A further feature of the invention resides in the provision of a clutch 
aggregate or assembly for use with a combustion engine, particularly in a 
motor vehicle. Such clutch aggregate or assembly comprises a torque 
transmitting friction clutch including a pressure plate, a counterpressure 
plate (such as a flywheel) which is rotatable about a predetermined axis, 
means for non-rotatably connecting the pressure plate to the 
counterpressure plate with limited freedom of axial movement, a torque 
transmitting clutch disc between the two plates, at least one resilient 
device acting upon the pressure plate to bias the clutch disc against the 
counterpressure plate, friction linings forming part of the clutch disc 
and being engageable with and disengageable from at least one of the two 
plates and being subject to wear as a result of repeated engagement with 
and disengagement from the at least one plate, an adjusting unit including 
means for compensating at least for wear upon the friction linings to 
thereby maintain the bias of the at least one resilient device upon the 
pressure plate at a substantially constant value during the useful life of 
the friction clutch, means for engaging and disengaging the friction 
clutch including actuating means movable along a predetermined path to 
disengage the friction clutch, means for gradually reducing the torque 
which can be transmitted by the friction clutch and/or by the clutch disc 
during a portion of movement of the actuating means in a direction to 
disengage the friction clutch, and axially elastic means for coupling the 
friction clutch with an output shaft of the combustion engine. The 
coupling means has a stiffness or rigidity which is selected in such a way 
that any axial, turning, wobbling (tilting) and/or flexing vibrations 
which are induced by the output shaft of the engine and would normally be 
transmitted to the friction clutch are damped and/or otherwise suppressed 
by the coupling means to an extent which ensures proper operation of the 
friction clutch, and especially proper operation of the adjusting unit. 
The stiffness of the coupling means can be selected in such a way that the 
force to be applied to the actuating means for disengagement of the 
friction clutch is taken up by the coupling means without appreciable 
axial shifting of the clutch aggregate or assembly. 
The adjusting unit of the just discussed aggregate or assembly can comprise 
resilient means in series with the at least one resilient device. Such 
aggregate or assembly can further comprise means for damping rotational 
and/or axial and/or radial vibrations of the counterpressure plate and 
such damping means is connectable between the output shaft of the engine 
and the counterpressure plate. 
An additional feature of the invention resides in the provision of a 
driving unit, particularly for use in motor vehicles, which comprises an 
at least partially automatic (i.e., fully automatic or semiautomatic) 
transmission, an engine, and a torque transmitting friction clutch 
disposed between the engine and the transmission and being controlled at 
least in dependency on the operation of the transmission. The friction 
clutch of such driving unit comprises a pressure plate, a counterpressure 
plate (such as a flywheel) rotatable about a predetermined axis and 
connectable to the output shaft of the engine, means for non-rotatably 
connecting the pressure plate to the counterpressure plate with limited 
freedom of movement in the direction of the predetermined axis, a torque 
transmitting clutch disc between the two plates, at least one resilient 
device acting upon the pressure plate to bias the clutch disc against the 
counterpressure plate, friction linings forming part of the clutch disc 
and being engageable with and disengageable from at least one of the two 
plates and being subject to wear as a result of repeated engagement with 
and disengagement from the at least one plate, an adjusting unit including 
means for compensating at least for wear upon the friction linings to 
thereby maintain the bias of the at least one resilient device upon the 
pressure plate at a substantially constant value, means for engaging and 
disengaging the friction clutch including actuating means movable along a 
predetermined path to engage and disengage the friction clutch, and torque 
varying means including means for gradually reducing the torque which can 
be transmitted by the friction clutch and/or by the clutch disc during a 
portion of movement of the actuating means along the predetermined path in 
a direction to disengage the friction clutch. 
The at least one resilient device (e.g., a diaphragm spring) can have a 
degressive force-to-displacement ratio, at least during a portion of 
movement of the actuating means along the predetermined path in a 
direction to disengage the friction clutch. 
The means for varying the torque which can be transmitted by the friction 
clutch can include means for reducing the transmittable torque during 
approximately 40-70 percent of movement of the actuating means along the 
predetermined path in a direction to disengage the friction clutch and for 
gradually increasing the torque which can be transmitted by the friction 
clutch during approximately 40-70 percent of movement of the actuating 
means along the path in a direction to engage the friction clutch. 
The novel features which are considered as characteristic of the invention 
are set forth in particular in the appended claims. The improved friction 
clutch itself, however, both as to its construction and its mode of 
operation, together with additional features and advantages thereof, will 
be best understood upon perusal of the following detailed description of 
certain presently preferred specific embodiments with reference to the 
accompanying drawings.

DESCRIPTION OF PREFERRED EMBODIMENTS 
Referring first to FIGS. 1 and 2, there is shown a torque transmitting 
friction clutch 1 which comprises a housing or cover 2 and a pressure 
plate 3 which is non-rotatably but axially movably (within limits) 
connected to the cover 2. A resilient device in the form of a diaphragm 
spring 4 is installed, in stressed condition, between the bottom wall or 
end wall 2a of the cover 2 and the pressure plate 3 so as to bias the 
pressure plate in a direction to the left, as viewed in FIG. 2, namely 
against the adjacent set of friction linings 7 forming part of a torque 
transmitting clutch plate or clutch disc 8. The diaphragm spring 4 is 
tiltable relative to the cover 2 at a location which is determined by an 
annular seat assembly 5 (hereinafter called seat for short) carried by the 
bottom wall 2a. The normal stressed condition of the diaphragm spring 4 is 
such that it urges the pressure plate 3 against the adjacent set of 
friction linings 7 of the clutch disc 8 whereby a second set of friction 
linings forming part of the clutch disc bears against the adjacent 
friction surface of a rotary counterpressure plate 6 here shown as a 
flywheel and hereinafter called flywheel for short. The illustrated clutch 
disc 8 comprises a centrally located hub 8a which can be installed on the 
input shaft (not shown) of a variable-speed transmission in a motor 
vehicle and carries two sets of friction linings 7 with resilient segments 
10 between the two sets of linings. The clutch 1 is engaged and the 
flywheel 6 transmits torque to the input shaft of the transmission when 
the two sets of friction linings 7 of the clutch disc 8 are clamped 
between the neighboring friction surfaces of the pressure plate 3 and 
flywheel 6. 
The means for connecting the pressure plate 3 with the cover 2 comprises 
several circumferentially extending leaf springs 9 (see particularly FIG. 
1) which hold the pressure plate 3 against rotation but permit it to move, 
within limits, axially of the cover 2. The purpose of the resilient 
segments 10 between the two sets of friction linings 7 forming part of the 
clutch disc 8 is to establish a progressive buildup (variation) of torque 
during engagement of the friction clutch 1. Such resilient segments permit 
limited axial shifting of the two sets of friction linings 7 toward each 
other to thus establish a progressive increase of axial forces which act 
upon the friction linings. However, it is equally within the purview of 
the invention to employ a clutch disc which replaces the illustrated 
clutch disc 8 and comprises one or two sets of friction linings 7 having 
no freedom of axial movement relative to each other. Such friction linings 
can be glued or otherwise affixed to opposite sides of a suitable 
plate-like carrier surrounding the hub of the modified clutch disc. 
The illustrated diaphragm spring 4 comprises a circumferentially complete 
annular main or primary or basic portion 4a which is adjacent one or more 
axially extending protuberances or portions 3a of the pressure plate 3 and 
serves to generate the major part of forces which are necessary to bias 
the pressure plate 3 against the adjacent friction linings 7 so as to urge 
the other set of friction linings against the friction surface of the 
flywheel 6. The main or primary portion 4a of the diaphragm spring 4 
carries radially inwardly extending yieldable prongs 4b having radially 
innermost portions or tips 4c engageable by a bearing, a pedal or other 
component which forms part of means for disengaging the friction clutch 1. 
The main or primary portion 4a of the diaphragm spring 4 includes a 
radially outer part which engages the portion or portions 3a of the 
pressure plate 3, and a radially inner part which is disposed between two 
portions 11 and 12 of the seat 5. Such radially inner portion of the main 
or primary portion 4a is tiltable between the portions 11 and 12 in order 
to move the radially outer portion of the diaphragm spring 4 toward or 
away from the flywheel 6, i.e., to engage or disengage the clutch 1. 
The illustrated portions 11 and 12 of the seat 5 are wire rings which flank 
the radially inner part of the main or primary portion 4a of the diaphragm 
spring 4 at a location radially inwardly of the portion or portions 3a of 
the pressure plate 3. A resilient distance or displacement monitoring 
sensor 13, here shown as a diaphragm spring, is provided to bias the inner 
ring or portion 11 of the seat 5 toward the bottom wall 2a of the cover 2. 
The illustrated resilient sensor 13 in the form of a diaphragm spring can 
be replaced by other biasing means without departing from the spirit of 
the invention. The annular radially outer portion 13b of the sensor 13 is 
circumferentially complete and comprises an outermost part or portion 13a 
which reacts against an abutment 14 at the inner side of the cover 2. The 
sensor 13 further comprises radially inwardly extending elastic tongues 
13c which bear upon the adjacent side of the ring 11. 
The abutment 14 for the radially outermost part 13b of the annular portion 
13a of the sensor 13 can comprise a circumferentially complete annulus 
which is welded or otherwise secured to the inner side of the cover 2. 
Alternatively, the abutment 14 can comprise two or more arcuate sections 
which are affixed to the inner side of the cover 2 to be engaged by the 
adjacent part 13b of the sensor 13. The individual segments or portions of 
the abutment 14 can be bonded, riveted or otherwise affixed to the cover 
2. It is also possible to provide an abutment 14 consisting of one or more 
projections which are of one piece with the cover 2 and are configurated 
to extend into the path of leftward movement (reference being had to FIG. 
2) of the radially outermost part 13a of the sensor 13. Such abutment can 
consist of radially inwardly deformed portions of the cover 2 or it can 
include one or more lugs or prongs which are bent from the adjacent 
portion of the cover to thus establish holes in the cover adjacent the 
radially outermost part 13a of the sensor 13. The making of such inwardly 
extending portions, lugs or prongs can take place subsequent to 
installation of the sensor 13 in the interior of the cover 2. The sensor 
13 can be stressed as a result of the making of abutment 14, or the sensor 
is already maintained in stressed condition at the time the abutment 14 is 
either installed or formed as an integral part of the cover 2. 
It is also possible to provide a more pronounced connection between the 
sensor 13 and the abutment 14. For example, it is possible to provide a 
bayonet mount which can establish a positive but separable connection 
between the sensor 13 and the abutment 14. The configuration of the 
bayonet mount can be such that the sensor 13 can be installed in the cover 
2 to have its radially outermost part 13a located to the left of the 
abutment 14, as viewed in FIG. 2. The radially outermost part 13a is then 
shifted axially toward the bottom wall 2a of the cover 2 to stress the 
sensor 13 and to cause such radially outermost part 13a to advance over 
the adjacent portion or portions of the abutment 14. The next step 
involves turning of the sensor 13 relative to the abutment 14 and/or vice 
versa so as to releasably lock the radially outermost part 13a of the 
sensor in the position which is shown in FIG. 2, namely at the right-hand 
side of the abutment 14. In such friction clutches, the radially outermost 
part 13a of the sensor 13 can comprise a plurality of prongs or arms 
extending radially outwardly beyond the circumferentially complete annular 
portion 13b of the sensor. The abutment 14 then comprises a plurality of 
recesses or tooth spaces, which permit the prongs of the sensor 13 to pass 
therethrough before the sensor is turned so that its prongs are moved out 
of alignment with the tooth spaces of the abutment 14. 
The means for centering the diaphragm spring 4 and/or the sensor 13 in the 
cover 2 and for simultaneously preventing rotation of such springs 
relative to the pressure plate 3 comprises axially parallel rivets 15. 
Such rivets can further serve as a means for centering the rings 11 and 12 
of the seat 5 in the cover 2. Each rivet 15 comprises an elongated shank 
15a which extends in parallelism with the axis (X--X) of rotation of the 
clutch plate 8 and is anchored in the bottom wall 2a of the cover 2. The 
shanks 15a extend through slots 4d between the neighboring prongs 4b of 
the diaphragm spring 4. The tongues 13c of the sensor 13 comprise portions 
13d which straddle the adjacent portions of the shanks 15a so that the 
rivets 15 hold the sensor 13 against rotation in the cover 2. 
The resilient sensor 13 is designed to furnish a substantially constant 
force during a predetermined stage of its axial deformation. The purpose 
of the sensor 13 is to bias the ring 11 toward the bottom wall 2a of the 
cover 2 as well as to take up the clutch disengaging force when such force 
is being applied to the tips 4c of prongs 4b forming part of the diaphragm 
spring 4. Depression of the tips 4c in a direction to the left, as viewed 
in FIG. 2, results in tilting of the diaphragm spring 4 between the rings 
11 and 12 of the seat 5 whereby the main or primary portion 4a of the 
spring 4 is moved away from the flywheel 6 so that the pressure plate 3 
can be retracted by the leaf springs 9 and releases the adjacent set of 
friction linings 7. In other words, the clutch disc 8 ceases to rotate 
with the flywheel 6. The arrangement is such that a state of equilibrium, 
or a state at least closely approximating equilibrium, exists between (a) 
that force which is generated upon the ring 11 during the application of 
disengaging force to the tips 4c of the prongs 4b and (b) the counterforce 
which is furnished by the sensor 13 and acts upon the ring 11. The term 
"disengaging force" is intended to denote that maximum force which must be 
applied to the tips 4c of the prongs 4b in order to disengage the friction 
clutch 1. Such disengaging force can also be applied to the prongs 4b by 
disengaging levers or by a pedal, not shown. 
In accordance with a feature of the invention, the ring 12 between the 
diaphragm spring 4 and the bottom wall 2a of the cover 2 is biased by an 
adjusting or regulating unit 16 which is installed in the cover. The 
adjusting unit 16 ensures that, when the rings 11 and 12 of the seat 5 are 
shifted axially in a direction toward the pressure plate 3 and flywheel 6, 
no undesirable clearance will develop between the ring 12 and the cover 2. 
Otherwise stated, there will be no clearance between the ring 12 and the 
diaphragm spring 4. Such positioning of the ring 12 relative to the 
diaphragm spring 4 is desirable and advantageous because this ensures that 
there is no undesirable lost motion during actuation of the friction 
clutch 1 which, in turn, ensures optimum efficiency and superior operation 
of the friction clutch. Axial shifting of the rings 11 and 12 of the seat 
5 toward the pressure plate 3 and flywheel 6 will take place as a result 
of wear upon the friction surfaces of the pressure plate 3 and flywheel 6 
as well as (and particularly) due to wear upon the friction linings 7. The 
exact mode of automatic operation of the adjusting unit 16 will be 
described in full detail with reference to the diagrams which are shown in 
FIGS. 8, 9, 10 and 11. 
The adjusting unit 16 comprises a spring-biased ring-shaped adjusting or 
wear compensating member 17 which is shown in FIGS. 3 and 4. This 
adjusting member 17 is installed between the diaphragm spring 4 and the 
bottom wall 2a of the cover 2 and comprises a set of inclines or ramps 18. 
All of the ramps 18 are inclined in the same direction circumferentially 
of the member 17. When the member 17 is installed in the cover 2, its 
ramps 18 face the bottom wall 2a. That side of the member 17 which faces 
away from the bottom wall 2a is flat or substantially flat and is provided 
with a circumferentially extending groove 19 (FIG. 2) which receives a 
portion of the ring 12. In this manner, the ring 12 (which is centered by 
the shanks 15a of the rivets 15) centers the member 17 in the cover 2. The 
configuration of the groove 19 and/or of the adjacent portion of the 
member 17 can be such that the ring 12 is not only held against 
uncontrolled radial movements but is also held against axial movement 
relative to the member 17. For example, the configuration of the surface 
bounding the groove 19 can be such that the ring 12 can be received 
therein by snap action. Alternatively, the plane surface of the member 17 
which faces away from the bottom wall 2a can be provided with spaced-apart 
projections or other configurations which enable the member 17 to 
clampingly or otherwise engage (e.g., by snap action) the adjacent 
portions of the ring 12 and to thus ensure that this ring is held against 
any uncontrolled radial and/or axial movements relative to the member 17. 
If the temperature of the friction clutch 1 in the region of the adjusting 
unit 16 fluctuates within a wide range, it is advisable to provide for 
some compensatory movement between the ring 12 and the member 17 of the 
adjusting unit 16. For example, this can be achieved by making the ring 12 
a split ring so that it can expand or contract in the circumferential 
direction of the member 17. It is also possible to assemble the ring 12 of 
two, three or more discrete arcuate sections, i.e., to provide two or more 
interruptions in such ring in order to even more fully compensate for 
eventual pronounced fluctuations of the temperature of the adjusting unit 
16. This enables the ring 12 to conform its diameter to the varying 
diameter of the groove 19. 
The member 17 which is shown in FIGS. 3 and 4 is made of a plastic 
material, for example, of a heat-resistant thermoplastic substance which 
can be reinforced by glass fibers or the like. This renders it possible to 
mass-produce the member 17 in an injection molding or other suitable 
machine. However, it is equally within the purview of the invention to 
make the member 17 of a metallic sheet material or of a sintered metal. 
Still further, it is within the scope of the invention to make the ring 12 
of one piece with the member 17. This is possible regardless of whether 
the member 17 is made of a metallic or plastic material. Analogously, the 
ring 11 can be made of one piece with the sensor 13; all that is necessary 
is to provide the tongues 13c of the sensor 13 with suitable projections 
in the form of beads or the like which together constitute a composite or 
one-piece ring 11. 
The rivets 15, and more particularly the shanks 15a of such rivets, 
preferably further constitute a means for centering the member 17 of the 
adjusting unit 16 in the cover 2 of the friction clutch 1. The rivets 15 
are preferably equidistant from each other in the circumferential 
direction of the cover 2. The shanks 15a extend through suitable openings 
21 which are provided in the member 17 and are bounded by surfaces 20 
which are engaged by the shanks 15a to thus center the member 17 in the 
cover 2. The illustrated openings 21 are elongated slots having a 
substantially constant width (as measured in the radial direction) and 
extend circumferentially of the member 17. These openings 21 are closely 
adjacent the radially inner portion of the member 17. As can be seen in 
FIG. 3, the member 17 further comprises lobes 22 which are disposed 
radially inwardly of the respective openings 21. 
The member 17 which is shown in FIG. 3 comprises three openings 21 and a 
total of five ramps 18 between each pair of neighboring openings. The 
slopes (note the angle 23 in FIG. 4) of the ramps 18 are selected in such 
a way that the ramps enable the member 17 to compensate for wear upon the 
pressure plate 3, flywheel 6 and friction linings 7 during the entire 
useful life of the friction clutch 1. The same applies for the length of 
the openings 21 in the circumferential direction of the member 17. Such 
length is selected with a view to permit an angular adjustment of the 
member 17 relative to the bottom wall 2a of the cover 2 which is necessary 
for compensation for the aforediscussed wear upon the pressure plate 3, 
flywheel 6 and friction linings 7. The length of the openings 21 can be 
selected in such a way that the member 17 is free to perform an angular 
movement in the range of between 8.degree. and 60.degree., preferably 
within a range of between 10.degree. and 30.degree.. In the embodiment 
which is illustrated in FIG. 3, the angular adjustability of the member 17 
relative to the cover 2 is approximately 12.degree.. Furthermore, the 
angle 23 (which is shown in FIG. 4 and denotes the slope of the ramps 18) 
is also in the range of 12.degree.. This angle 23 is selected in such a 
way that, when the ramps 18 of the member 17 and the complementary 
inclines or ramps 24 of a second annular displacing member 25 (shown in 
FIGS. 5 and 6 and hereinafter called annulus) are in frictional engagement 
with each other, the member 17 and the annulus 25 cannot slip because the 
friction between the abutting surfaces of the ramps 18 and 24 is too 
pronounced. Depending on the nature of the material of the member 17 and 
annulus 25 and on the finish of the abutting surfaces of the ramps 18 and 
24, the angle 23 can be in the range of between 5.degree. and 20.degree.. 
The member 17 is stressed in the circumferential direction by a ring-shaped 
torsion spring 26 which is shown in FIGS. 1, 2, 7 and 7a. The bias of the 
spring 26 is selected in such a way that the member 17 is stressed in a 
direction which is necessary for adjustment in order to compensate for 
wear upon the pressure plate 3, flywheel 6 and friction linings 7. In 
other words, the spring 26 tends to bias the member 17 in a direction such 
that, as the ramps 18 slide along the complementary ramps 24 of the 
annulus 25, this results in axial displacement of the member 17 in a 
direction toward the pressure plate 3, i.e., axially of and away from the 
bottom wall 2a of the cover 2. It is clear that the illustrated torsion 
spring 26 constitutes but one form of means for biasing the member 17 in a 
direction to slide along the annulus 25 and to thereby advance axially 
toward the pressure plate 3. This torsion spring comprises a relatively 
small number of convolutions 35 (for example, not more than two 
convolutions) and two legs 27 and 28. The leg 27 extends radially 
outwardly (see particularly FIGS. 7 and 7a), and the leg 28 extends in 
part radially and in part axially (see FIG. 7a). The leg 27 is 
non-rotatably anchored in or is otherwise connected with the member 17, 
and the leg 28 is non-rotatably anchored in or is otherwise secured to the 
cover 2. The spring 26 is installed in stressed condition. 
A presently preferred form of the displacing annulus 25 is shown in FIGS. 5 
and 6. This annulus comprises the aforediscussed ramps 24 which are 
complementary to the ramps 18 of the member 17. The surfaces along which 
the ramps 18 abut the ramps 24 can be congruent surfaces. The angle 29 
which is shown in FIG. 6 preferably matches the angle 23 which is shown in 
FIG. 4. As can be readily seen by comparing FIGS. 3 and 5, the 
distribution of ramps 24 on the annulus 25 is the same as, or at least 
similar to, that of the ramps 18 on the member 17. The annulus 25 is 
non-rotatably secured to the housing 2. To this end, the annulus 25 is 
provided with a plurality of holes 30 which can receive portions of the 
rivets 15 so that such rivets also serve as a means for non-rotatably 
coupling the annulus 25 to the bottom wall 2a of the cover 2. This can be 
seen in the upper portion of FIG. 2. 
FIG. 2 further shows, by broken lines, that the means for biasing the 
member 17 in the circumferential direction of the cover 2 can comprise an 
additional torsion spring 26a which can be configurated in the same way as 
the torsion spring 26. Thus, one leg of the torsion spring 26a can be 
anchored in the member 17 and its other leg can be anchored in the cover 
2. The torsion spring 26a is also installed in stressed condition so that 
it always tends to turn the member 17 relative to the cover 2. 
An advantage of the utilization of two torsion springs 26, 26a is that 
their bias can increase under the action of centrifugal force when the 
friction clutch 1 is in use and its cover 2 rotates with the pressure 
plate 3 and flywheel 6. The flywheel 6 can receive torque from the output 
element of an engine in a motor vehicle. For example, the increased bias 
of the spring 26 in response to the action of centrifugal force can be 
compensated for by the torsion spring 26a. To this end, the springs 26 and 
26a are convoluted in such a way that, at least when acted upon by 
centrifugal force, they generate and apply to the member 17 forces which 
act in opposite directions as seen in the circumferential direction of the 
member 17. The diameters of convolutions of the torsion spring 26a are 
larger than the diameters of convolutions 35 of the torsion spring 26. 
Reference may be had to FIG. 2. This enables the designer of the clutch to 
select the centrifugal forces acting upon the torsion springs 26 and 26a 
in such a way that the forces acting upon the member 17 in the 
circumferential direction are at least substantially balanced. Adequate 
balancing can be achieved by appropriate selection of the diameters of 
convolutions of the springs 26, 26a, by appropriate selection of the 
diameters of wires of which these springs are made and/or by appropriate 
selection of the number of their convolutions. FIG. 2 shows that the 
torsion spring 26 is located radially inwardly and the torsion spring 26a 
is located radially outwardly of the member 17. However, it is equally 
possible to install each of these springs radially inwardly or radially 
outwardly of the member 17. 
FIG. 7 shows the torsion spring 26 in a plan view. When this spring is not 
under stress, its legs 27, 28 make an angle 31 which can be in the range 
of 40.degree.-120.degree.. The leg 27 will be moved (relative to the leg 
28) to the position 32 when the friction linings 7 are new (i.e., prior to 
being subjected to any wear). The leg 27 assumes the position 33 of FIG. 7 
when the linings 7 have undergone a maximum permissible amount of wear. 
The angle 34 of adjustment (between the positions 32 and 33 shown in FIG. 
7) is approximately 12.degree.. The spring 26 of FIG. 7 is designed in 
such a way that, when in unstressed condition, only a single convolution 
35 extends between the legs 27 and 28. The remaining portion of the spring 
26 (namely outside of the angle 31) has two convolutions 35 (FIG. 7a) 
which overlie each other as seen in the axial direction of the spring 26. 
The spring 26a is similar to the spring 26 but, in the embodiment of FIGS. 
1 and 2, has a larger diameter and is stressed in a different direction as 
concerns its bias upon the member 17. The force which the spring 26 
applies to the member 17 is greater than the force of the spring 26a. 
When the wear upon the component parts of the friction clutch 1 is minimal, 
i.e., when the clutch is yet to be put to use, the angular positions of 
the member 17 and annulus 25 relative to each other are such that the 
axially extending peaks 18a of the ramps 18 forming part of the member 17 
extend close to or actually abut the axially extending peaks 24a of the 
ramps 24 on the annulus 25. In other words, the combined thickness of the 
member 17 and annulus 25 then assumes a minimum value, i.e., these parts 
occupy a minimum amount of space in the axial direction of the cover 2 
between the bottom wall 2a and the diaphragm spring 4. 
In the friction clutch 1 of FIGS. 1 and 2, the annulus 25 constitutes a 
separately produced part which is installed at the inner side of the 
bottom wall 2a of the cover 2. However, it is also possible to make the 
annulus 25 an integral part of the cover 2; for example, the lobes 24 can 
be stamped out of the bottom wall 2a to extend toward the member 17 of the 
adjusting unit 16. Such mode of making the annulus 25 (namely its lobes 
24) is particularly advantageous if the cover 2 is made of a single piece 
of metallic sheet material. 
The shoulders 38 on the pallets 36 of lobes 22 of the member 17 can be 
utilized to ensure proper angular positioning of the member 17 in the 
cover 2 during assembly of the friction clutch 1. The shoulders 38 can be 
engaged by a suitable turning or retaining tool which reacts against the 
cover 2. The tool is put to use during assembly of the friction clutch 1 
and is removed when the attachment of the cover 2 to the flywheel 6 (by 
threaded fasteners 6a one of which is shown in FIG. 2) is completed. The 
adjusting unit 16 becomes operative as soon as the aforementioned tool is 
removed, i.e., as soon as the member 17 is free to turn relative to the 
cover 2 (if and when necessary) to compensate for wear upon the parts 3, 6 
and/or 7. As shown in FIG. 1, the bottom wall 2a of the cover 2 has 
circumferentially extending elongated slot-shaped windows 37 which enable 
the prongs or analogous extensions of the tool to engage the shoulders 38 
on the pallets 36 of the lobes 22 of the member 17 during assembly of the 
friction clutch 1. The shoulders 38 can be replaced with other 
configurations (e.g., holes) in or on the member 17, as long as the tool 
can properly engage and hold the member 17 in requisite position during 
assembly of the friction clutch 1. The length of the windows 37 should at 
least suffice to ensure that the member 17 can be turned back through the 
maximum angle which is required to compensate for wear upon the pressure 
plate 3, flywheel 6 and/or friction linings 7. It is also possible to 
assemble the friction clutch 1 in a first step and to thereupon employ a 
tool which is to be used to turn the member 17 relative to the cover 2. 
The prongs of the tool are inserted through the windows 37 of the bottom 
wall 2a and engage the shoulders 38 on the pallets 36 of the lobes 22. The 
member 17 is then turned back in a direction to ensure that its ramps 18 
cooperate with the ramps 24 of the annulus 25 in a sense to move the 
member 17 closer to the bottom wall 2a to a position from which the member 
17 must turn in order to compensate for wear upon the parts 3, 6 and/or 7 
in actual use of the friction clutch 1. The member 17 is then located at a 
minimum distance from the bottom wall 2a and is secured in such position, 
for example, with a clamp or a pin extending into registering openings of 
the cover 2 and member 17 to prevent angular displacement of the thus 
coupled parts 2, 17 relative to each other. The clamp or pin is removed 
when the attachment of the cover 2 to the flywheel 6 is completed, i.e., 
the unit 16 is then ready to perform its adjusting action if and when 
necessary, depending on the extent of wear upon the pressure plate 3, 
flywheel 6 and/or friction linings 7. 
The dimensions of the windows 37 in the cover 2 are selected in such a way 
that the member 17 can be returned to its "retracted" position (at a 
minimal distance from the bottom wall 2a) if and when the cover 2 is to be 
detached from the flywheel 6. This involves disengagement of the clutch 1 
(i.e., the application of axial force against the tips 4c of the prongs 4b 
in a direction toward the clutch disc 8) so that the diaphragm spring 4 no 
longer exerts an axially oriented force against the ring 11 of the seat 5 
and the member 17 can be readily turned relative to the cover 2. 
Referring to the diagram of FIG. 8, the sinusoidal curve 40 denotes the 
axially oriented force which develops in response to changes of conicity 
of the diaphragm spring 4 as a result of deformation between two abutments 
spaced apart from each other a distance corresponding to that of the seat 
5 from the projecting portion or portions 3a of the pressure plate 3. The 
distance between such abutments is measured along the abscissa, and the 
force which is generated by the diaphragm spring 4 is measured along the 
ordinate of the coordinate system of FIG. 8. The (operating) point 41 of 
the curve 40 denotes the force which is generated by the diaphragm spring 
4 upon installation of the friction clutch 1 and while the clutch is 
engaged; at such time, the spring 4 exerts a maximum force upon the 
portion or portions 3a of the pressure plate 3 and the latter exerts a 
maximum force which is used to clamp the friction linings 7 of the clutch 
disc 8 between the friction surfaces of the pressure plate 3 and flywheel 
6. The point 41 can be shifted along the curve 40 toward or away from the 
abscissa by changing the conicity of the diaphragm spring 4 in assembled 
condition of the friction clutch 1. 
The curve 42 denotes in FIG. 8 the axial spreading force which is applied 
by the resilient segments 10 between the two sets of friction linings 7. 
Such spreading force of the segments 10 opposes the force which the 
diaphragm spring 4 applies to the pressure plate 3. It is desirable and 
advantageous that the force which develops as a result of resilient 
deformation of the segments 10 at least match the bias of the diaphragm 
spring 4; it is also possible to select the mounting of the diaphragm 
spring 4 and the resiliency and bias of the segments 10 in such a way that 
the force which is denoted by the curve 42 exceeds the force which is 
denoted by the curve 40. The stressing of the resilient segments 10 
decreases in response to disengagement of the friction clutch 1, and the 
extent to which the stressing decreases is denoted by the distance 43. 
This results in a corresponding axial shifting or deformation of the 
diaphragm spring 4 whereby the segments 10 assist the disengagement of the 
friction clutch. In other words, the required maximum disengaging force is 
less than that which would be necessary at the point 41 of the curve 40 in 
FIG. 8 if the resilient segments 8 were omitted. The point 44 on the curve 
40 denotes the magnitude of the force of diaphragm spring 4 at the instant 
of disengagement of the friction clutch 1, i.e., the friction linings 7 
are no longer engaged by the friction surfaces of the pressure plate 3 and 
flywheel 6 when the point 44 is exceeded. Due to the degressive 
characteristic curve of the diaphragm spring 4, the disengaging force 
which is to be applied at such time is much less than that corresponding 
to the force denoted by the point 41 of the curve 40. The disengaging 
force which must be applied in the friction clutch 1 decreases all the way 
to the minimum or lowest point 45 of the sinusoidal curve 40. From there 
on, the required disengaging force rises again and the extent of axial 
movement of the tips 4c of prongs 4b along their predetermined path can be 
selected in such a way that the magnitude of this force does not exceed 
that at the point 44 (i.e., the maximum disengaging force) and preferably 
remains therebelow. In other words, the force should not rise above that 
denoted by the point 46. 
The magnitude of force which is generated by the sensor 13 is denoted by 
the curve 47 which is shown in FIG. 9. This curve actually denotes the 
force which is generated when the conicity of the sensor 13 is changed as 
a result of stressing. Such change in stressing of the sensor 13 takes 
place as a result of variations of the distance between two abutments 
whose radial spacing corresponds to that of the abutment 14 at the inner 
side of the cover 2 from the ring 11 of the seat 5. The distance 48 
covered by the sensor 13 is that during which the axial force generated by 
the sensor remains substantially constant. The magnitude of this force is 
selected in such a way that it at least approximates the magnitude of the 
clutch disengaging force as denoted by the point 44 on the curve 40 of 
FIG. 8. The supporting force to be furnished by the sensor 13 is less than 
that at the point 44 of the curve 40 by a value corresponding to the lever 
arm of the diaphragm spring 4. In most instances, such transmission ratio 
is between 1:3 and 1:5 but can also be less than 1:3 or greater than 1:5 
for certain applications of the improved friction clutch. 
The just mentioned transmission ratio of the diaphragm spring 4 denotes the 
ratio of radial distance of the seat 5 from the portion or portions 3a of 
the pressure plate 3 to the radial distance of the seat 5 from the tips 4c 
of the prongs 4b forming part of the diaphragm spring 4 and being 
depressible, for example, by a disengaging bearing of the friction clutch. 
The mounting of the sensor 13 in the friction clutch 1 is selected in such 
a way that the sensor can perform an axial movement in the region of the 
seat 5, namely in a direction toward the friction linings 7, to an extent 
corresponding at least to the axial adjustment of the pressure plate 3 
toward the flywheel 6 as a result of wear upon the friction surfaces of 
the parts 3, 6 and as a result of wear upon the friction linings 7. This 
ensures that the axially oriented supporting force for the seat 5 remains 
constant regardless of the wear upon the parts 3, 6 and 7. In other words, 
the substantially linear portion 48 of the curve 47 in FIG. 9 should have 
a length not less than that corresponding to the aforediscussed extent of 
wear and preferably exceeding the latter. This ensures that the adjusting 
unit 16 can also compensate, at least in part, for eventual tolerances 
during assembly of the friction clutch 1. 
In order to ensure the establishment of a practically unchanged (i.e., 
predetermined) release point 44 for the friction linings 7 when the 
friction clutch 1 is disengaged, it is possible to employ torque varying 
means 10 known as a so-called twin-segment biasing means which is to 
operate between the two sets of friction linings 7. Such biasing means can 
comprise pairs of discrete parallel resilient segments which are disposed 
back-to-back. The segments which are disposed back-to-back can be 
subjected to a certain initial stress in the axial direction of the clutch 
disc 8 so that the axial force which is generated by all such pairs of 
segments relative to each other at least matches the disengaging force 
denoted by the point 44 of the sinusoidal curve 40 when the clutch disc 8 
is not clamped between the pressure plate 3 and the flywheel 6. It is 
preferred to ensure that the combined force of the pairs of resilient 
segments slightly exceed that force of the diaphragm spring 4 which is 
denoted by the point 44 on the curve 40 of FIG. 8. Prestressing of 
resilient segments between the friction linings 7 of the clutch disc 8 
renders it possible to at least substantially compensate for so-called 
penetration or embedding losses which develop during the useful life of 
the friction clutch 1 as a result of penetration of the segments 10 into 
the adjacent friction linings 7. Such penetration of segments 10 into the 
adjacent linings 7 is to be expected in actual use of the friction clutch 
1. It has been found that an axial stressing or give of the segments 10 in 
the range of 0.3 mm to 0.8 mm (preferably approximately 0.5 mm) is quite 
satisfactory. By properly limiting the extent of axial movability of the 
two sets of friction linings 7 relative to each other and by properly 
selecting the bias of the segments 10 between the two sets of friction 
linings, one can ensure that, at least during disengagement of the 
friction clutch 1, the pressure plate 3 covers a predetermined distance 43 
in a direction away from the friction linings under the action of the 
resilient segments 10. In order to achieve such predetermined distance 43, 
it is possible to limit the extent of axial movement of the two sets of 
friction linings 7 in directions toward as well as away from each other, 
e.g., by the provision of suitable stops, i.e., in directions to stress 
the segments 10 as well as to enable these segments to dissipate energy. 
Suitable resilient means for use between the two sets of friction linings 
7 are disclosed, for example, in commonly owned copending German patent 
application Serial No. P 42 06 880.0. 
In order to guarantee an optimal operation of the friction clutch 1, i.e., 
to ensure that the adjusting unit 16 will be capable of automatically 
compensating for wear upon the parts 3, 6 and/or 7 of the friction clutch, 
it is desirable to ensure that the sum of forces which are applied by the 
diaphragm spring 4, sensor 13 and resilient segments 10, as well as the 
force which is applied to the diaphragm spring 4 solely by the sensor 13 
when the pressure plate 3 is already disengaged from the adjacent friction 
linings 7, at least equals but preferably exceeds the variable disengaging 
force which is being applied to the tips 4c of prongs 4b during 
disengagement of the friction clutch 1. The variable disengaging force is 
denoted by the curve 49 in the diagram of FIG. 10. 
The heretofore discussed mode of operation of the friction clutch 1 
pertains primarily or exclusively to a predetermined mode of installing 
the diaphragm spring 4 and without taking into consideration the wear upon 
the friction linings 7. When a certain amount of wear has taken place 
(such wear is particularly pronounced upon the friction linings 7), the 
position of the pressure plate 3 changes in that the pressure plate 
migrates toward the flywheel 6 whereby the conicity of the diaphragm 
spring 4 (and hence the bias of this spring upon the portion or portions 
3a of the pressure plate) changes accordingly while the friction clutch 1 
remains in engaged condition. Such change of bias of the diaphragm spring 
4 upon the pressure plate 3 entails that the point 41 of the curve 40 in 
the diagram of FIG. 8 migrates toward the point 41' and that the point 44 
migrates toward the point 44'. This terminates the state of equilibrium 
between the diaphragm spring 4 and the sensor 13 at the ring 11 during 
disengagement of the friction clutch 1. Wear upon the friction linings 7 
entails an increase in the magnitude of force which is applied by the 
diaphragm spring 4 to the sensor 13 and also causes a shifting of the 
progress of the disengaging force in a sense toward an increase of such 
force. The thus obtained progress of the disengaging force is denoted in 
FIG. 10 by the broken-line curve 50. Since the magnitude of the 
disengaging force increases, the axially oriented force of the sensor 13 
upon the diaphragm spring 4 during disengagement of the friction clutch 1 
is overcome so that the sensor 13 yields in the region of the seat 5 
through an axial distance corresponding essentially to the extent of wear 
upon the friction linings 7. During such deformation stage of the sensor 
13 (which can be said to constitute a means for monitoring the extent of 
wear upon the parts 3, 6 and/or 7), the diaphragm spring 4 bears against 
the portion or portions 3a of the pressure plate 3 whereby the conicity of 
the spring 4 changes together with the amount of energy which is stored 
therein. Thus, the energy which is stored by the diaphragm spring 4 also 
changes together with the force which the spring 4 exerts upon the ring 
11, i.e., upon the sensor 13 and upon the pressure plate 3. As can be seen 
in FIG. 8, such change takes place in a sense to reduce the magnitude of 
the force which is applied by the diaphragm spring 4 and continues to take 
place until the magnitude of the axial force applied by the spring 4 to 
the sensor 13 at the ring 11 is at least substantially neutralized or 
balanced by the oppositely directed force which is exerted by the sensor 
13. In other words, and referring again to the diagram of FIG. 8, the 
points 41' and 44' of the curve 40 then migrate toward the points 41 and 
44, respectively. When the reestablishment of the state of equilibrium is 
completed, the pressure plate 3 is again ready to be disengaged from the 
adjacent friction linings 7. During the aforediscussed stage of adjustment 
in order to compensate for wear upon the friction linings 7, while the 
friction clutch 1 is being disengaged, the member 17 of the adjusting unit 
16 is caused to turn about the axis X--X of the clutch disc 8 under the 
bias of the stressed torsion spring 26 which causes a displacement of the 
ring 12 to an extent corresponding to the extent of wear upon the friction 
linings 7; this, in turn, eliminates any play at the seat 5. When the 
adjusting step is completed, the magnitude of the disengaging force again 
corresponds to that denoted by the curve 49 in the diagram of FIG. 10. The 
curves 50A and 51 in the diagram of FIG. 10 denote the axial displacement 
of the pressure plate 3 when the magnitude of the disengaging force varies 
in accordance with the curves 49 and 50, respectively. 
The curves which are shown in the diagram of FIG. 11 denote the variations 
of forces acting upon the cover 2 and upon the sensor 13 during 
disengagement of the friction clutch 1. The extreme values are omitted. 
Starting with the engaged condition of FIG. 2, the cover 2 and the 
pressure plate 3 are first acted upon by a force whose magnitude 
corresponds to the operating or installation point 41 of the diaphragm 
spring 4 as denoted by the curve 40 of FIG. 8. As the disengagement of the 
friction clutch 1 progresses, the magnitude of the axial force exerted by 
the diaphragm spring 4 upon the cover 2 and the ring 12 decreases in 
accordance with the curve 52 of FIG. 11, namely to the point 53. When the 
point 53 is exceeded in the direction of disengagement of the friction 
clutch 1, a conventional frictional clutch (wherein the diaphragm spring 
is tiltable at a fixed location relative to the clutch cover, i.e., 
wherein the ring 11 is fixedly installed in the cover) would operate in 
such a way that the force exerted by the diaphragm spring 4 upon the cover 
2 at the level of the seat 5 would change (reverse) its direction. 
However, the novel friction clutch 1 operates in such a way that the 
change in the axial direction of the force applied by the diaphragm spring 
4 in the region of the seat 5 is taken up by the sensor 13. When the 
magnitude of the force which is being applied by the diaphragm spring 4 
reaches the value denoted by the point 54 on the curve 52 of FIG. 11, the 
diaphragm spring 4 becomes disengaged from the portion or portions 3a of 
the pressure plate 3. The resilient segments 10 between the two sets of 
friction linings 7 generate an axially oriented force which assists the 
disengagement of the friction clutch 1 at least to the point 54 on the 
curve 52 of FIG. 11. The force which is generated by the resilient 
segments 10 decreases as the extent of displacement of tips 4c of prongs 
4b toward the clutch disc 8 increases during disengagement of the friction 
clutch 1, i.e., in response to progressing axial displacement of the 
pressure plate 3 in a direction away from the flywheel 6. Thus, the curve 
52 of FIG. 11 denotes a resultant of a disengaging force which is being 
applied to the tips 4c during disengagement of the friction clutch 1 on 
the one hand and of the axial force which is being applied by the 
resilient segments 10 of the clutch disc 8 upon the diaphragm spring 4 in 
the region of portion or portions 3a of the pressure plate 3. When the 
point 54 is exceeded in the direction of disengagement of the friction 
clutch 1, the axially oriented force which is being applied by the 
diaphragm spring 4 to the ring 11 is compensated for by the oppositely 
directed force which is being applied by the sensor 13. These two forces 
are balanced by the pressure plate 3 not later than when the axial 
pressure upon the friction linings 7 is terminated. As the disengaging 
operation progresses, the axially oriented force which is being applied by 
the sensor 13 at the seat 5 preferably exceeds, at least slightly, the 
prevailing disengaging force. The portion 55 of the curve 52 in the 
diaphragm of FIG. 11 indicates that, as the extent of movement to 
disengage the friction clutch 1 increases, the disengaging force (and the 
force applied by diaphragm spring 4 to the ring 11) decreases when 
compared with the disengaging force denoted by the point 54 of the curve 
52. The broken-line curve 56 in the diagram of FIG. 11 denotes that 
condition of the friction clutch 1 when the friction linings 7 have 
undergone a certain amount of wear but prior to any compensation for such 
wear in the region of the seat 5. It will be noted that the change of 
orientation (conicity) of the diaphragm spring 4 due to wear upon the 
friction linings 7 results in an increase of the magnitude of forces which 
are being applied to the cover 2, to the ring 11 and/or to the sensor 13. 
This causes the point 54 to migrate in a direction toward 54' which, in 
turn, entails that in the course of the next-following disengaging 
operation the axial force which is being applied by the diaphragm spring 4 
to the sensor 13 at the ring 11 exceeds the oppositely directed force of 
the sensor 13; this causes an adjustment in the aforedescribed manner as a 
result of axial relaxation of the sensor 13. Such adjustment entails that 
the point 54' migrates toward the point 54 which, in turn, reestablishes 
the desired state of equilibrium at the seat 5, namely between the 
diaphragm spring 4 and the sensor 13. 
In actual practice (i.e., when the friction clutch 1 is in use), 
adjustments by the unit 16 are effected continuously or nearly 
continuously (i.e., by minute steps). The distances between the various 
points on the curves of FIGS. 8 to 11 are greatly exaggerated for the sake 
of clarity. 
It is very likely that certain changes of various functional parameters 
and/or operating points will take place during the useful life of the 
friction clutch 1. For example, improper actuation of the friction clutch 
1 can result in overheating of the resilient segments 10 in the clutch 
disc 8 which can cause a reduction of the resiliency of these segments, 
i.e., a reduction of the extent of axial movability of the parts 10. 
Nevertheless, it is possible to ensure reliable operation of the friction 
clutch 1 by appropriate selection of the characteristic curve 40 of the 
diaphragm spring 4 and a corresponding conformance of the curve 47 
denoting the displacement-to-force relationship of the sensor 13. A 
reduction of axial movability of the segments 10 would merely entail that 
the conicity of the diaphragm spring 4 in the friction clutch 1 of FIGS. 1 
and 2 would change in a sense to reduce the magnitude of the force which 
the spring 4 exerts upon the portion or portions 3a of the pressure plate 
3. This can be seen in the diagram of FIG. 8. Furthermore, this would 
bring about a corresponding change of axial deformation of the sensor 13 
and a corresponding axial displacement of the ring 11. 
In accordance with a further feature of the invention, it is possible to 
construct the improved friction clutch in such a way that the resultant of 
forces acting upon the diaphragm spring 4 increases in response to 
increasing wear upon the friction linings 7. Such increase can be limited 
to a certain stage or portion of the maximum permissible displacement due 
to wear upon the friction linings 7. As mentioned above, the wear upon the 
friction linings is normally more pronounced than the wear upon the 
flywheel 6 and upon the pressure plate 3; therefore, the preceding and the 
next-following passages of this description refer primarily or exclusively 
to wear upon the friction linings. The increase of the magnitude of forces 
acting upon the diaphragm spring 4 can take place as a result of 
appropriate design of the sensor 13. FIG. 9 shows by broken lines, as at 
47a, the characteristic curve denoting a thus modified sensor 13 within 
the range 48. If the magnitude of forces acting upon the diaphragm spring 
4 increases in response to progressing wear upon the friction linings 7, 
one can at least partially compensate for a reduction of the force which 
the spring 4 applies to the pressure plate 3 due to a reduction of 
resiliency of the segments 10, e.g., as a result of penetration or 
embedding of these segments into the adjacent friction linings 7. It is 
particularly advantageous if the force for the diaphragm spring 4 
increases proportionally with (i.e., at the same rate or nearly at the 
same rate as) the setting or reduction of bias of the segments 10, for 
example, due to the aforediscussed embedding into the adjacent friction 
linings 7. In other words, as the thickness of the clutch disc 8 in the 
region of the friction linings 7 decreases (i.e., as the distance between 
the two sets of friction linings decreases due to the reduced bias of the 
segments 10 as a result of penetration into the friction linings and/or 
due to wear upon the friction linings), the magnitude of forces acting 
upon the diaphragm spring 4 increases accordingly. It is of particular 
advantage if the magnitude of such forces increases in such a way that the 
increase is more pronounced during a first stage and less pronounced 
during a next-following second stage. These two stages are within the 
distance 48 as measured along the abscissa of the coordinate system which 
is shown in FIG. 9. The just outlined design is desirable and advantageous 
because the major part of penetration of segments 10 into the adjacent 
friction linings 7 takes place mainly during a relatively short period of 
the full useful life of the friction clutch; thereafter, the positions of 
the segments 10 relative to the adjacent friction linings 7 are more or 
less stabilized. Thus, once a certain penetration has taken place, this 
variable parameter or factor can be disregarded because it no longer 
affects the operation of the adjusting unit 16. The change of magnitude of 
the force acting upon the diaphragm spring 4 can also take place at least 
during a certain stage of wear upon the friction linings 7. 
The preceding description of operation of the adjusting unit 16 to 
compensate for wear upon the friction linings 7 did not take into 
consideration the axially oriented forces which are or which can be 
generated by the leaf springs 9 serving to axially movably but 
non-rotatably couple the pressure plate 3 to the flywheel 6 and cover 2. 
If the leaf springs 9 are installed in stressed condition so that they 
tend to move the pressure plate 3 axially and away from the adjacent 
friction linings 7, i.e., in a sense to bias the portion or portions 3a of 
the pressure plate 3 against the diaphragm spring 4, the leaf springs 9 
are in a condition to assist the disengagement of the friction clutch 1. 
Thus, the axially oriented force which is applied by the leaf springs 9 is 
superimposed upon the forces which are being applied by the sensor 13 and 
by the diaphragm spring 4 as well as upon the disengaging force which is 
being applied (e.g., by a suitable disengaging bearing) against the tips 
4c of the prongs 4b. Such function of the leaf springs 9 is not considered 
in the diagrams of FIGS. 8 to 11. The overall force which is being applied 
to the diaphragm spring 4 in disengaged condition of the friction clutch 1 
to cause the spring 4 to bear upon the ring 12 of the seat 5 is the sum of 
forces which are generated primarily by the leaf springs 9, sensor 13 and 
the applied disengaging force acting upon the tips 4c of prongs 4b forming 
part of the spring 4. 
The leaf springs 9 can be installed between the cover 2 and the pressure 
plate 3 in such a way that their axially oriented force acting upon the 
diaphragm spring 4 increases in response to progressing wear upon the 
friction linings 7. For example, the magnitude of axial force exerted by 
the leaf springs 9 upon the diaphragm spring 4 in response to increasing 
wear upon the friction linings 7 can increase in accordance with a curve 
47b which is shown in the diagram of FIG. 9 and denotes the variations of 
such force upon the spring 4 within the distance 48. FIG. 9 further shows 
that, as the deformation of the sensor 13 increases, the restoring force 
of the leaf springs 9 upon the pressure plate 3 (this force is also 
applied to the diaphragm spring 4) also increases. By totalizing the 
forces denoted by the curve 47b and the characteristic curve of the 
diaphragm spring, one arrives at a resultant force which acts upon the 
spring 4 in the axial direction in a sense to bias this spring against the 
ring 12 of the seat 5. In order to obtain a variation of forces as denoted 
by the curve 47b of FIG. 9, it is desirable to design the sensor 13 in 
such a way that its characteristic curve corresponds to that shown at 47c 
in FIG. 9. By summarizing the forces denoted by the curves 47b and 47c in 
the diagram of FIG. 9, one arrives at a sum of forces denoted by the curve 
47a. Thus, the magnitude of the force to be applied by the sensor 13 can 
be reduced by the simple expedient of stressing the leaf springs 9. 
Furthermore, by properly designing and mounting the leaf springs 9, it is 
possible to reduce (at least in part) the bias of the resilient segments 
10 and/or (at least in part) the extent of penetration of segments 10 into 
the adjacent friction linings 7. Thus, one can ensure that the diaphragm 
spring 4 maintains a substantially unchanged operating point or the same 
operating range, i.e., the bias of the spring 4 upon the pressure plate 3 
remains at least substantially unchanged during the entire useful life of 
the friction clutch 1. It is further necessary or desirable to take into 
consideration (during designing of the improved friction clutch and 
particularly in connection with the design of the sensor 13 and leaf 
springs 9) the resultant axial forces which are generated by the torsion 
springs 26, 26a and act upon the member 17 of the adjusting unit 16 in a 
sense to oppose the bias of the sensor 13 and/or the bias of the leaf 
springs 9. 
If the friction clutch of the present invention is designed to employ 
prestressed leaf springs 9, it is further necessary or advisable to take 
into consideration that the prestressing of the springs 9 influences the 
axial force which the pressure plate 3 applies to the adjacent friction 
linings 7. Thus, if the leaf springs 9 are prestressed in a sense to urge 
the pressure plate 3 toward the diaphragm spring 4, the force which is 
applied by the diaphragm spring 4 is reduced by the extent of prestressing 
of the leaf springs 9. Consequently, the friction clutch then operates in 
such a way that the resultant axial force acting upon the pressure plate 3 
and hence upon the adjacent friction linings 7 includes the force of the 
spring 4 and the force resulting from prestressing of the leaf springs 9. 
If one assumes that the curve 40 in the diagram of FIG. 8 denotes the 
resultant of the forces due to bias of the spring 4 plus the force 
attributable to prestressing of the leaf springs 9 in unused condition of 
the friction clutch, a reduction of the distance of the pressure plate 3 
from the flywheel 6 due to wear upon the friction linings 7 would result 
in a shifting of the resulting forces in a sense toward a reduction of 
forces. FIG. 8 shows a broken-line curve 40a which corresponds, for 
example, to a wear in the range of 1.5 mm. Such wear can develop during 
the useful life of the friction clutch 1, and a shifting from the curve 40 
toward the curve 40a results in a reduction of axial force which is being 
applied by the diaphragm spring 4 to the sensor 13 during disengagement of 
the friction clutch; such reduction of the axial force is attributable to 
the fact that, as the wear upon the friction linings 7 progresses, the 
moment which is being applied by the leaf springs 9 to the spring 4 and 
acts in the opposite direction also increases. Such moment develops due to 
the existence of a radial clearance between the seat 5 and the diameter of 
the annulus defined by the portion or portions 3a of the pressure plate 3, 
i.e., at the locus of engagement between the pressure plate and the spring 
4. In designing the friction clutch 1, it is of particular importance to 
ensure that the increasing bias of the leaf springs 9 (as a result of wear 
upon the friction linings 7) is less than the increase of disengaging 
force which is also attributable to wear upon the friction linings and 
causes a tilting of the sensor 13 which is necessary to cause the unit 16 
to carry out the necessary adjustment. Otherwise, the biasing force of the 
pressure plate 3 upon the friction linings 7 would decrease in engaged 
condition of the friction clutch and this would prevent any adjustments of 
the seat 5. 
An important advantage of the torque varying resilient segments 10 is that 
the torque which is being transmitted by the hub 8a of the clutch disc 8 
to the input element of a transmission decreases gradually, at least 
during a portion of movement of the prongs 4b of the diaphragm spring 4 in 
the direction to disengage the friction clutch 1. Furthermore, the 
resilient segments 10 ensure a gradual (progressive) increase of torque 
which is being transmitted by the clutch disc 8, at least during a portion 
of movement of the prongs 4b in a direction to engage the friction clutch 
1, particularly during the initial stage of clamping of the two sets of 
friction linings 7 at opposite sides of the resilient segments 10 between 
the friction surfaces of the pressure plate 3 and flywheel or 
counterpressure plate 6. 
Another important advantage of the improved friction clutch 1 is that the 
stressing of the diaphragm spring 4 in engaged condition of the friction 
clutch 1 remains at least substantially unchanged during each and every 
stage of useful life of the friction clutch. Otherwise stated, the bias of 
the diaphragm spring 4 upon the adjacent portion or portions 3a of the 
pressure plate 3 remains at least substantially unchanged irrespective of 
the extent of wear upon the friction linings 7. 
A further important advantage of the friction clutch 1 is that, due to 
gradual reduction of torque which can be transmitted by the clutch disc 8 
during disengagement of the friction clutch, it is now possible to greatly 
reduce (minimize) the magnitude of the force which must be applied to 
disengage the clutch, i.e., it is possible to optimize the progress of 
variation of disengaging force when the pressure plate 3 is being moved 
away from the flywheel 6. This is due to the fact that the resilient 
segments 10 assist the actuation of the friction clutch, particularly the 
disengagement of the clutch when the tips 4c of the prongs 4b forming part 
of the diaphragm spring 4 move in a direction to permit the resilient 
segments 10 to dissipate energy while the pressure plate 3 is in the 
process of moving axially and away from the flywheel 6. Though FIGS. 1 and 
2 show a friction clutch wherein the resilient segments 10 are disposed 
between the two sets of friction linings 7, it is equally possible to 
employ resilient means which are or is analogous to the segments 10 but 
are or is designed and mounted to apply a reaction force to the prongs 4b 
of the diaphragm spring 4 (i.e., to the actuating means of the means for 
engaging and disengaging the friction clutch 1) and/or to another part of 
the diaphragm spring 4 and/or to the pressure plate 3 and/or to the 
flywheel 6. Such reaction force acts counter to the action of the 
diaphragm spring 4 in a direction to urge the pressure plate 3 against the 
adjacent set of friction linings 7. Furthermore, the resilient segments 10 
and/or their equivalent(s) is or are disposed in series with the diaphragm 
spring 4. 
A particularly desirable and advantageous feature of the resilient segments 
10 and/or of their equivalent(s) is that they can ensure a gradual 
reduction of torque which is being transmitted by the clutch disc 8 during 
a portion of movement of the prongs 4b in a direction to effect a 
disengagement of the friction clutch 1, i.e., that such reduction of 
transmittable torque takes place during a certain stage of movement of the 
pressure plate 3 away from the flywheel 6. 
An equivalent of the resilient segments 10 can be installed in the power 
train between the seat 5 for the diaphragm spring 4 and the fasteners 6a 
which secure the cover 2 to the flywheel 6, or between the diaphragm 
spring 4 and the fasteners 6a. Alternatively, or in addition to the 
provision of torque varying means between the two sets of friction linings 
7 and/or between the seat 5 and the fasteners 6a, it is also possible (and 
often desirable and advantageous) to install torque varying means between 
the seat 5 or the diaphragm spring 4 on the one hand, and the friction 
surface of the pressure plate 3 on the other hand. Reference may be had, 
for example, to published German patent application Serial No. 37 42 354 
and/or to published German patent application Serial No. 1 450 201. 
Still further, it is possible to install an equivalent of the resilient 
segments 10 in another portion of the friction clutch in addition to the 
segments 10 between the two sets of friction linings 7. As concerns the 
installation of resilient segments between two sets of friction linings, 
reference may be had, for example, to published German patent application 
Serial No. 36 31 863. 
Still further, it is within the purview of the invention to gradually 
increase and/or reduce the magnitude of transmitted torque during 
engagement or disengagement of the friction clutch by installing an 
equivalent of resilient segments 10 in a manner as disclosed in published 
German patent application Serial No. 21 64 297. Thus, it is possible to 
employ a composite (twin-mass) flywheel including a first mass which is 
connected to and receives torque from the output element of an internal 
combustion engine, and a second mass which constitutes or includes a 
counterpressure plate and is axially movably coupled with the first mass 
by resilient means so that the resilient means opposes a movement of the 
second mass at least toward or at least away from the first mass. 
The operation of the improved friction clutch is particularly satisfactory 
if the resilient torque varying means is installed to permit movements of 
certain parts of the friction clutch toward and away from each other 
against the opposition of a spring bias. The arrangement is preferably 
such that the magnitude of the force opposing axial movements of certain 
parts relative to each other is smallest when the friction clutch is 
disengaged but that the magnitude of the force opposing axial movements of 
certain parts relative to each other gradually rises to a maximum value 
during clamping of the friction linings 7, i.e., during engagement of the 
friction clutch. Such rise of the opposing force need not take place 
during the entire engagement stage. It has been found that the improved 
friction clutch operates highly satisfactorily if the magnitude of the 
aforediscussed force which opposes axial movements of certain parts of the 
friction clutch relative to each other gradually increases during between 
40 and 70 percent of movement of the prongs 4b in a direction to engage 
the friction clutch and gradually decreases during between about 40 and 70 
percent of movement of the prongs 4b in a direction to disengage the 
friction clutch. The remaining portions of movement of the prongs 4b 
during engagement and disengagement of the friction clutch are needed to 
ensure reliable interruption of power flow, to establish full transmission 
of torque and/or to compensate for possible deformation of certain parts 
of the friction clutch (especially the clutch disc, the pressure plate 
and/or the counterpressure plate). 
The feature that the diaphragm spring 4 has a degressive 
force-to-displacement ratio is desirable and advantageous because this 
renders it possible to minimize the forces which are required to engage or 
disengage the friction clutch, especially to minimize the forces which are 
required to disengage the clutch. Such degressive character need not be 
effective during the entire stage of disengagement of the friction clutch. 
Otherwise stated, it is desirable to ensure that the magnitude of the 
force which is being applied by the diaphragm spring 4 will decrease at 
least during a certain stage of its compression or deforming movement 
while the friction clutch is being disengaged so that, during such stage, 
the stressing and/or deformation of the diaphragm spring 4 is assisted by 
the resilient torque varying segments 10 and/or their equivalents. 
At the same time, and due to the degressive force-to-distance ratio of the 
spring 4 during a certain stage of disengagement of the friction clutch, 
the magnitude of the force which the spring 4 exerts upon the friction 
linings 7 through the pressure plate 3 is on the decrease. In the absence 
of any other superimposed spring-generated forces, the effective force 
which is required to disengage the improved friction clutch equals the 
difference between the force which is being applied by the torque varying 
means 10 (and/or their equivalent or equivalents) and the force of the 
diaphragm spring 4. When the pressure plate 3 is being lifted off the 
adjacent friction linings 7, i.e., when the pressure plate releases the 
clutch disc 8, the remainder of movement of the prongs 4b in a direction 
to complete the disengagement of the friction clutch 1 will be effected 
primarily by the diaphragm spring 4. The force-to-displacement ratios of 
the diaphragm spring 4 and torque varying means 10 can be related to each 
other in such a way that, when the clutch disc 8 is released by the 
pressure plate 3, a relatively small force is needed to actuate the 
diaphragm spring. In other words, by properly relating the aforediscussed 
ratios or characteristics of the spring 4 and torque varying means 10 
(e.g., by causing these characteristics to be identical or to only 
negligibly deviate from each other), it is possible to ensure that only a 
very small force (and in an extreme case zero force) is necessary to move 
the diaphragm spring 4 up to the instant of disengagement of the clutch 
disc 8 by the pressure plate 3. 
Though it is possible to bias the pressure plate 3 by a resilient device 
other than a diaphragm spring 4, it is presently preferred to employ a 
diaphragm spring and to mount the diaphragm spring in the seat 5 so that 
it is tiltable relative to the cover 2 and can bear against one or more 
selected portions 3a of the pressure plate 3. This also simplifies the 
means for engaging and disengaging the friction clutch 1 because the 
radially inwardly extending prongs 4b of such diaphragm spring can 
constitute the actuating means of such clutch engaging/disengaging means. 
However, it is equally possible to employ modified engaging/disengaging 
means, e.g., including levers which are pivotably mounted on the cover 2 
or on another part of the friction clutch. 
The diaphragm spring 4 can be replaced, for example, with coil springs 
which are then installed in the friction clutch in such a way that their 
force acting axially upon the pressure plate 3 reaches a maximum value 
when the clutch is engaged but decreases during disengagement of the 
clutch. These characteristics can be achieved, for example, by mounting 
the coil springs in such a way that they are inclined with reference to 
the rotational axis X--X of the clutch. 
The utilization of a diaphragm spring 4 which is tiltable relative to a 
seat 5 on the housing 2 of the improved friction clutch is desirable and 
advantageous on the additional ground that this renders it possible to 
design the clutch as a so-called push-type or depression-type clutch. In 
such friction clutches, the means for disengaging the clutch is normally 
moved in a direction toward the pressure plate, i.e., in a direction to 
the left as seen in FIG. 2. However, the present invention can be embodied 
with equal advantage in so-called pull-type friction clutches wherein 
(again referring to FIG. 2) the prongs 4b or the equivalents of such 
actuating means must be moved in a direction to the right in order to 
disengage the friction clutch. 
As already described with reference to FIG. 8, it is often preferred to 
employ a diaphragm spring 4 having a substantially sinusoidal 
characteristic curve and being installed in the housing 2 in such a way 
that, when the friction clutch is engaged, the operating point of the 
diaphragm spring is located within a degressive portion of the curve 
following the first maximum of such curve. The so-called force ratio of 
the diaphragm spring 4 whose characteristic curve coincides with or is 
similar to the sinusoidal curve in the diaphragm of FIG. 8 can be within a 
range of approximately 1:0.4 and 1:0.7 between the first maximum and the 
next-following lowest point or minimum of the curve. 
If the improved friction clutch is installed in a motor vehicle, the means 
for engaging and disengaging can further comprise a pedal which resembles 
or is analogous to a standard gas pedal and is installed in the vehicle to 
serve as a means for moving the tips 4c of the prongs 4b along their 
predetermined path in a direction to disengage and/or engage the friction 
clutch. The utilization of a pedal which is similar or analogous to a gas 
pedal is of particular advantage in view of the aforediscussed 
characteristics of the improved friction clutch. Thus, and since the force 
which is required to disengage the friction clutch is relatively small or 
extremely small, such relatively small force can be selected (metered) 
with a high degree of accuracy and reproducibility if the means for moving 
the prongs 4b along their path is a pedal, i.e., a device whose 
manipulation is familiar to all drivers. 
A further important advantage of the improved friction clutch is that, due 
to the aforediscussed possibility of greatly reducing the maximum forces 
which must be applied during the entire useful life of the clutch (i.e., 
that the forces to be applied need not be increased as the wear upon the 
friction linings 7 progresses), it is now possible to reduce the 
dimensions of various component parts of the friction clutch and to 
greatly reduce the strength or stability of such parts. This, in turn, 
contributes to a significant reduction of the cost of the friction clutch. 
Furthermore, the aforediscussed reduction of disengaging force renders it 
possible to greatly reduce losses due to friction and/or losses due to 
decreasing resiliency of many parts of the friction clutch, particularly 
in the disengaging means. This greatly enhances the efficiency of the 
friction clutch disengaging system. Moreover, this renders it possible to 
achieve an optimal design of the friction clutch and to render the 
manipulation of the friction clutch more comfortable to the operator. 
The aforediscussed improvements which were described in connection with and 
are shown in the embodiment of FIGS. 1 to 11 can be utilized in a number 
of presently known friction clutches. Examples of friction clutches whose 
operation and/or other characteristics can be improved by incorporating 
therein the features of the present invention are those described and 
shown, for example, in German Pats. Nos. 29 16 755 and 29 20 932, in 
published German patent applications Serial Nos. 35 18 781 and 40 92 382, 
in published French patent applications Serial Nos. 2 605 692, 2 606 477, 
2 599 444 and 2 599 446, in British Pat. No. 1 567 019, in U.S. Pat. Nos. 
4,924,991, 4,191,285 and 4,057,131, in published Japanese patent 
application Serial No. 51-126452, and in Japanese Utility Models Nos. 
3-25026, 3-123, 2-124326, 1-163218, 3-19131 and 3-53628. 
Reference may also be had to commonly owned copending German patent 
applications Serial Nos. P 42 07 528.9 and P 42 06 904.1. 
The provision of the adjusting unit 16, which compensates for wear upon at 
least one component (particularly the friction linings 7) of the improved 
friction clutch 1, brings about the additional advantage that it is now 
possible to optimize the design and the operation of the friction clutch, 
especially of the diaphragm spring 4 which is called upon to bias the 
pressure plate 3 against the adjacent set of friction linings 7 in engaged 
condition of the clutch. The diaphragm spring 4 can be designed in such a 
way that it is merely called upon to furnish only that force which is 
required for transmission of the desired torque, i.e., to clamp the 
pressure plate 3 against the clutch disc 8 only with a force which ensures 
that the clutch disc 8 can transmit requisite torque to the input element 
of a variable-speed transmission or the like. As mentioned above, it is 
not absolutely necessary to employ a resilient device in the form of a 
diaphragm spring, such as the diaphragm spring 4; it is also possible to 
employ two or more resilient devices such as a set of coil springs which 
can be distributed and oriented in a manner as already described 
hereinbefore. 
FIGS. 12 and 13 illustrate certain details of a modified torque 
transmitting friction clutch 101. One of the differences between the 
friction clutches 1 and 101 is that the latter employs three coil springs 
126 (two shown in FIG. 12) which replace the torsion springs 26, 26a in 
the friction clutch 1 and serve to bias the annular member 117 of the 
adjusting unit 116 for the bias of the diaphragm spring 4. As concerns its 
function, the member 117 is an equivalent of the member 17, i.e., it can 
cooperate with an annulus corresponding to the annulus 25 of FIGS. 5 and 6 
to move the portion 111 of the seat 105 in a direction to the right (as 
viewed in FIG. 13) to an extent which is necessary to compensate for wear 
upon the friction linings 107 of the clutch plate or disc 108. 
The friction clutch 101 employs three coil springs 126 which are 
equidistant from each other in the circumferential direction of the 
housing or cover 2 and are installed in stressed condition to bias the 
member 117 relative to the bottom wall 2a of the cover. As can be seen in 
FIG. 14, the inner marginal portion of the member 117 is provided with 
axially, radially and circumferentially extending projections 127 which 
serve as stops for the adjacent ends of the respective coil springs 126. 
The stops 127 are acted upon by the respective springs 126 in a sense to 
tend to turn the member 117 about the axis of the cover 2 in a direction 
to move the portion 111 of the seat 105 for the diaphragm spring 4 toward 
the pressure plate 103. The springs 126 have an arcuate shape because they 
are adjacent the convex outer sides of arcuate guides 129 forming part of 
or affixed to the member 117. The other end of each coil spring 126 is in 
engagement with a discrete post 128 which is anchored in the bottom wall 
2a of the cover 2. The illustrated posts 128 have external threads which 
mate with the threads of tapped bores provided therefor in the bottom wall 
2a. However, it is equally possible to replace the externally threaded 
posts 128 with integral projections in the form of lugs or the like which 
are obtained by displacing selected portions of the bottom wall 2a in a 
direction toward the pressure plate 103. Such making of posts or like 
parts which are of one piece with the bottom wall 2a is particularly 
advantageous and simple if the cover 2 is made of a metallic sheet 
material. 
The length of the arcuate guides 129 is preferably selected in such a way 
that they can adequately guide the respective coil springs 126 during each 
stage of angular displacement of the member 117 relative to the bottom 
wall 2a, i.e., during each stage of compensation for wear upon the 
friction linings 107, pressure plate 103 and/or counterpressure plate 106. 
The configuration of the guides 129 is such that they can properly prop 
the respective coil springs 126 from within (i.e., at the concave sides of 
the arcuate springs) as well as in the axial direction of the bottom wall 
2a. Each of the guides 129 can define an arcuate groove or channel which 
receives a portion of the respective coil spring 126 between the 
respective post 128 and the respective projection 127. This ensures highly 
predictable positioning of the coil springs 126 relative to the member 117 
and guarantees that these coil springs can turn the member 117 in the 
proper direction (to move the seat portion 111 toward the pressure plate 
103) whenever necessary in order to compensate for wear upon the linings 
107, pressure plate 103 and/or counterpressure plate 106. The 
configuration of the surfaces bounding the channels of the guides 129 on 
the member 117 can conform to the configuration of the adjacent portions 
of the respective coil springs 126. Such configuration of the surfaces 
bounding the channels or grooves in the guides 129 ensures that the coil 
springs 126 are adequately guided when the cover 2 is idle as well as when 
the cover is rotated by the counterpressure plate 106 (this 
counterpressure plate can constitute or form part of a flywheel which 
receives torque from the output shaft of a combustion engine in a motor 
vehicle). 
In order to even more reliably ensure optimal retention of coil springs 126 
in requisite positions relative to the bottom wall 2a and the member 117, 
the radially inner portion of the bottom wall 2a can be provided with 
axially extending arms 130 which are disposed radially inwardly of the 
coil springs (see FIG. 13). The individual arms 130 can be replaced with a 
circumferentially complete cylindrical collar of the bottom wall 2a. The 
arms 130 or the aforementioned circumferentially complete collar of the 
bottom wall 2a can perform the additional function of serving as an 
abutment for the adjacent portions of the diaphragm spring 4, i.e., such 
collar or the arms 130 can limit the extent of dissipation of energy by 
the diaphragm spring 4. 
The provision of means for guiding the coil springs 126 exhibits the 
advantage that, when the friction clutch 101 is rotated by a combustion 
engine or the like, the convolutions of the springs 126 cannot leave the 
illustrated positions under the action of centrifugal force, i.e., they 
cannot move into frictional engagement with the adjacent portions (such as 
ramps) of the member 117; this would result in the development of 
undesirable friction which would prevent the springs 126 from changing the 
angular position of the member 117 in a manner to accurately compensate 
for wear upon the friction linings 107, pressure plate 103 and/or 
counterpressure plate 106. When the friction clutch 101 is driven, the 
coil springs 126 preferably behave not unlike solid bodies, i.e., they are 
in frictional engagement with the adjacent guides 129 and such frictional 
engagement suffices to prevent any angular displacement of the member 117. 
The arrangement can be such that, when the rotational speed of the 
friction clutch 101 exceeds the idling speed of the engine, frictional 
engagement between the coil springs 126 and the guides 129 under the 
action of centrifugal force suffices to prevent any angular displacement 
of the member 117 relative to the bottom wall 2a of the cover 2, i.e., the 
springs 126 cannot change the angular position of the member 117. Thus, 
the angular position of the member 117 with reference to the cover 2 (in 
order to move the seat portion 111 toward the pressure plate 103) can take 
place only when the rotational speed of the friction clutch 101 does not 
exceed the idling speed of the engine. In other words, it is necessary to 
operate the friction clutch 101 in such a way that its rotational speed is 
relatively low in order to enable the springs 126 to change the angular 
position of the member 117 relative to the bottom wall 2a (if necessary). 
It is equally possible to block any turning of the member 117 relative to 
the bottom wall 2a in any one of a number of other ways, i.e., not 
necessarily as a result of pronounced frictional engagement with the 
surfaces bounding the grooves or channels of the respective guides 129. 
For example, the arrangement may be such that the coil springs 126 can 
change the angular position of the member 117 relative to the bottom wall 
2a only when the friction clutch 101 is not driven. 
The just discussed feature of the friction clutch 101 can be incorporated 
with equal advantage in the friction clutch 1 of FIGS. 1 and 2. The 
arrangement may be such that the angular position of the member 17 
relative to the annulus 25 can be changed only when the clutch 1 is not 
driven at all or when the clutch 1 rotates within a relatively low range 
of speeds. For example, the housing or cover 2 of the friction clutch 1 of 
FIGS. 1 and 2 can be provided with means which prevent the torsion spring 
26 and/or 26a from changing the angular position of the member 17 relative 
to the annulus 25 when the member 17 is acted upon by centrifugal force, 
i.e., when the friction clutch 1 is driven by the engine in a motor 
vehicle or the like. For example, the bottom wall 2a of the cover 2 in the 
friction clutch 1 can carry one or more flyweights which move radially 
outwardly under the action of centrifugal force to thereby interfere with 
any changes in the angular position of the member 17 relative to the 
annulus 25, either by directly engaging the member 17 and/or by preventing 
the spring 26 and/or 26a from changing the angular position of the member 
17 in the cover 2. The flyweight or flyweights can be designed and mounted 
to bear against the radially innermost portion of the member 17 when the 
friction clutch 1 of FIGS. 1 and 2 is driven. The flyweight or flyweights 
must be capable of engaging and holding the member 17 with a force which 
exceeds the bias of the springs 26, 26a, at least when the rotational 
speed of the friction clutch 1 reaches a certain value. 
Referring again to FIGS. 12-14, the friction clutch 101 can be modified by 
providing radial supports for portions of or for the entire coil springs 
126. Such radial supports can be installed on or they can form part of the 
bottom wall 2a of the cover 2 in the friction clutch 101; for example, the 
radial supports can be made of one piece with the posts 128. Thus, each 
post 128 can be replaced with a substantially L-shaped element which 
includes a portion extending in the circumferential direction of the cover 
2 and into the adjacent end convolutions of the respective coil spring 
126. Such portions of the L-shaped elements act not unlike retainers and 
hold the surrounding end convolutions of the respective coil springs 126 
against radial movement relative to the bottom wall 2a. 
FIG. 13 illustrates that the wire ring 11 of the seat 5 which is shown in 
FIG. 2 can be omitted. More specifically, the wire ring 11 is replaced by 
a radially inner portion 111 of the sensor 113. The portion 111 can be 
assembled of several sections each forming part of one of the tongues 113c 
of the sensor 113. Those sides of the tongues 113c which engage the 
diaphragm spring 4 in lieu of a wire ring 11 or the like can have a convex 
or substantially convex shape. Thus, the sensor 113 of FIG. 13 can perform 
the combined functions of the sensor 13 and wire ring 11 in the friction 
clutch 1 of FIGS. 1 and 2. 
FIGS. 15 to 17 illustrate certain details of a further torque transmitting 
friction clutch 201 wherein the circumferentially complete annular 
adjusting member 17 or 117 is replaced with a set of discrete button- or 
washer-like adjusting and wear compensating members 217. The discrete 
members 217 are equidistant from each other in the circumferential 
direction of the cover or housing 202 and each of these members has a ramp 
218 which extends at one of its sides in the circumferential direction to 
cooperate with an adjacent ramp 224 of the annulus 225 forming part of the 
bottom wall 202a of the cover 202. Each of the illustrated members 217 has 
a central opening 219 (e.g., a circular bore or hole) which receives a 
portion of an axially parallel pin-shaped extension 215a of a rivet in 
such a way that each member 217 can turn about the axis of the respective 
extension 215a. The annulus 225 is an integral part of the bottom wall 
202a and is provided with the aforementioned ramps 224 cooperating with 
the ramps 218 of the neighboring members 217 to automatically shift (when 
necessary) the ring 212 of the seat 205 toward the pressure plate 203 in 
order to compensate for wear upon the friction linings 207, the pressure 
plate 203 and/or the counterpressure plate (not shown in FIG. 15). The 
members 217 are turnable about the axes of the respective extensions 215a 
by springs 226 in a sense to move the ramps 218 along the neighboring 
ramps 224 and to thus shift the ring 212 toward the pressure plate 203. 
Each spring 226 resembles a helix which surrounds the respective extension 
215a, which reacts against the bottom wall 202a and which bears against 
the corresponding member 217. The end portions of the helical springs 226 
are suitably bent so that they can more reliably engage the bottom wall 
202a and the respective members 217, respectively. For example, the end 
portions of the springs 226 can be provided with lugs, legs or like 
projections. When the diaphragm spring 204 is moved axially of the 
friction clutch 201 of FIG. 15 due to wear upon the friction linings 207, 
pressure plate 203 and/or the non-illustrated counterpressure plate, the 
springs 226 are free to change the angular positions of the respective 
members 217 relative to the corresponding extensions 215a and to thus move 
the ring 212 toward the pressure plate 203; this compensates for the 
aforediscussed wear, primarily upon the friction linings 207. 
The sensor 213 of the friction clutch 201 of FIG. 15 bears against lugs 214 
which are shown in the form of integral portions of the axially extending 
part of the cover 202. The lugs 214 preferably constitute inwardly bent 
parts of the cover which are deformed to the extent necessary to engage 
the radially outer portion of the sensor 213. 
An advantage of the discrete annular members 217 is that they are less 
likely to change their positions under the action of centrifugal force, 
i.e., they are not likely to turn about the respective extensions 215a as 
a result of rotation of the cover 202 about its own axis. In other words, 
the adjusting action of such discrete members 217 is not affected by the 
magnitude of the centrifugal force. 
The discrete annular adjusting members 217 in the friction clutch which is 
shown in FIG. 15 can be replaced with discrete wedge-like or analogous 
adjusting members which are mounted for movement in the radial and/or 
circumferential direction of the cover 202 in order to cooperate with 
complementary parts on the bottom wall 202a in a sense to displace the 
ring 212 toward the pressure plate 203 when the need arises, i.e., in 
order to compensate for wear upon the counterpressure plate, the pressure 
plate 203 and/or the friction linings 207. Each wedge-like adjusting 
member can be provided with a longitudinally extending recess to receive a 
portion of an extension 215a or a like part of or on the bottom wall 202a. 
This ensures that each wedge-like member can carry out a movement only in 
a direction which is necessary to adjust the axial position of the ring 
212. The arrangement may be such that the wedge-like members which are to 
be used in lieu of the discrete washer-like members 217 of FIG. 15 are 
acted upon by centrifugal force in order to move radially and/or 
circumferentially of the wall 202a in order to compensate for wear, 
particularly for wear upon the friction linings 207. However, it is 
equally possible to employ springs which cooperate with wedge-like 
adjusting members to shift such adjusting members along suitable 
configurations (such as ramps 224) of the bottom wall 202a in order to 
move the ring 212 axially toward the pressure plate 203. The extensions 
215a can be replaced with other suitable guide means for the wedge-like 
adjusting members which can be used in lieu of the washer-like members 
217; for example, the bottom wall 202a can be grooved to establish 
predetermined paths for movement of the wedge-like members relative to the 
cover 202. 
The ramps 224 can be provided on the bottom wall 202a to project toward the 
adjacent annular members 217 from a plane which is normal to the axis of 
the cover 202. Alternatively, such ramps can be provided on the adjacent 
portions of the diaphragm spring 204. This also applies for the 
embodiments of FIGS. 1-2 and 12-13. It is also possible to provide the 
annular members 17, 117 and/or 217 with two sets of ramps 218, one at each 
side, and to provide complementary ramps 24, 124 or 224 on the bottom wall 
2a, 102a or 202a and on the corresponding diaphragm spring 4, 104 or 204. 
If the adjusting members are wedges or if they resemble wedges, it is 
advisable to make them from a lightweight material in order to minimize 
the influence of centrifugal force. 
The selection of materials for the cooperating ramps (such as 18 and 24) 
also plays an important role in connection with the reliability of 
adjustment of the diaphragm spring toward the clutch disc of the improved 
friction clutch. An important prerequisite is to select the material of 
the member 17 or 117 or of the members 217 and the material of the 
adjacent annulus 25, 125 or 225 in such a way that the ramps of such parts 
will not exhibit a tendency to adhere to each other irrespective of the 
momentary stage of the useful life of the respective friction clutch. For 
example, adherence of one set of ramps to the neighboring ramps can be 
prevented or avoided by coating at least one of these sets of ramps with a 
suitable friction-reducing material. If the one and/or the other set of 
ramps consists of a metallic material, the coating substance will or can 
be selected with a view to prevent corrosion. 
Another mode of preventing the ramps of one set from adhering to the ramps 
of the other set or sets (and from thus preventing, or interfering with 
accuracy of, adjustment of the diaphragm spring toward the pressure plate 
in order to compensate for wear) is to make the materials of the two or 
more sets of ramps (such as the materials of the annular member 17 and the 
annulus 25 in the embodiment of FIGS. 1 and 2) of materials having 
different thermal expansion coefficients. As a rule, the temperature of 
the friction clutch 1 will fluctuate in actual use as well as prior to and 
between actual use(s) or during and subsequent to actual use. This will 
entail certain minimal movements of the neighboring ramps 18, 24 relative 
to each other whenever the temperature of the member 17 and annulus 25 
changes. The aforementioned mode of selecting the materials of the member 
17 and annulus 25 (so that they have different thermal expansion 
coefficients) ensures that the ramps 18 cannot adhere to the ramps 24, 
i.e., that the adjusting unit 16 is always in condition to carry out all 
necessary adjustments in exact dependency on the extent of wear upon the 
friction linings 7 and/or pressure plate 3 and/or counterpressure plate or 
flywheel 6. 
Still another mode of preventing adherence of neighboring sets of ramps to 
each other is to select the configuration and/or the deformability 
(stability) of the corresponding parts (such as the member 17 and the 
annulus 25 in the friction clutch 1 of FIGS. 1 and 2) with a view to 
ensure that the action of centrifugal force upon the parts 17 and 25 is 
not the same, i.e., that such parts will perform certain movements 
relative to each other in response to rotation of the friction clutch 1 
with the result that the extent of movement of the ramps 18 will depart 
from that of the ramps 24 and the two sets of ramps will be incapable of 
adhering to one another. 
A further mode of preventing the ramps of one set from adhering to the 
ramps of the neighboring set or sets is to ensure that the ramps of at 
least one set (e.g., the ramps 18 in the friction clutch 1 of FIGS. 1-2) 
are caused to perform at least some axial movements relative to the 
neighboring ramps (such as 24) during each disengagement of the friction 
clutch (i.e., during movement of the tips 4c of prongs 4b of the diaphragm 
spring 4 along a predetermined path extending toward the pressure plate 3 
in order to deform the diaphragm spring 4 and to permit the leaf springs 9 
to shift the pressure plate 3 axially and away from the flywheel 6). The 
adjusting member 17 of the unit 16 in the embodiment of FIGS. 1-2 can be 
coupled with a suitable part or it can be provided with suitable parts 
which move axially in response to development of wear at 7, 3 and/or 6. 
Such part or parts can be installed adjacent the seat 5, e.g., on the 
diaphragm spring 4 and/or on the sensor 13. 
The diagram of FIG. 18 shows the characteristic curve 340 of a diaphragm 
spring corresponding to the diaphragm spring 4 of FIGS. 1 and 2. The curve 
340 has a minimum or lowest point 345 denoting a relatively small force 
which is generated by the diaphragm spring and is in the range of 
approximately 450 nm (as measured along the ordinate). The highest point 
or maximum of the curve 340 is located in the range of 7680 nm. The 
transmission of force as a result of deformation of the diaphragm spring, 
and as indicated by the curve 340 of FIG. 18, takes place while the 
diaphragm spring bears against one and reacts against another of two stops 
which are spaced apart from each other in the radial direction of the 
respective friction clutch. The situation is analogous to that described 
with reference to the characteristic curve 40 of the diaphragm spring 4 in 
the diagram of FIG. 8. 
The characteristic curve 340 of the diaphragm spring (such as 4) can be 
combined with the characteristic curve 342 of a resilient element 
corresponding to the segments 10 in the clutch disc 8 of the friction 
clutch 1. As can be seen in FIG. 18, the distance-to-force progress of the 
curve 342 is similar to that of the curve 340, i.e., these curves are 
rather close to each other which denotes that a friction clutch embodying 
the corresponding diaphragm spring and resilient segments 10 can be 
actuated in response to exertion of a very small force. Within the 
operating range of the resilient segments 10, the theoretical disengaging 
force corresponds to the difference between two vertically aligned points, 
one on the curve 340 and the other on the curve 342. One such difference 
is shown in FIG. 18, as at 360. The actually required disengaging force is 
further reduced by the corresponding lever arms of the actuating means, 
such as the prongs 4b of the diaphragm spring 4 in the friction clutch 1 
of FIGS. 1 and 2. All this is analogous to the construction and mode of 
operation of the friction clutch 1 as already described with reference to 
FIGS. 1-2 and 8-11. 
The diagram of FIG. 18 further contains a curve 440 which is indicated by 
broken lines and has a minimum or lowest point 445 denoting a negative 
force which is generated by a diaphragm spring. In other words, a certain 
part of the force which is denoted by the curve 440 does not assist in 
engagement of the friction clutch but rather tends to disengage the 
friction clutch. Thus, if the deformation of diaphragm spring which is 
denoted by the curve 440 progresses beyond the point 461, the friction 
clutch does not exhibit a tendency to become engaged but automatically 
remains disengaged. The broken-line curve 442 denotes in FIG. 18 the 
characteristic curve of resilient segments (such as 10 in the friction 
clutch of FIGS. 1-2) which can be used in conjunction with the diaphragm 
spring having a characteristic curve corresponding to that shown at 440. 
The curve 349 in the diagram of FIG. 19 denotes the progress of a 
disengaging force which is to be applied to the tips of the prongs of a 
diaphragm spring (i.e., to the actuating means of the means for engaging 
and disengaging the friction clutch) when the friction clutch employs a 
diaphragm spring and resilient segments of the character denoted by the 
curves 340 and 342 of FIG. 18. As can be seen in FIG. 19, the curve 349 
remains in the positive force range (above the abscissa of the coordinate 
system of FIG. 19) which means that a certain force in a direction to 
disengage the friction clutch must be applied as long as the friction 
clutch is to remain in disengaged condition (the pressure plate 3 of FIGS. 
1-2 is then disengaged from the adjacent set of friction linings 7). 
The broken-line curve 449 in the diagram of FIG. 19 denotes the progress of 
a clutch disengaging force which develops when the diaphragm spring and 
the resilient segments of the friction clutch exhibit characteristic 
curves of the type shown at 440 and 442 in the diagram of FIG. 18. The 
curve 449 includes a portion (at 449a) which denotes an initial decrease 
of the disengaging force toward the abscissa and thereupon transits from 
the positive side to the negative side of the abscissa. This denotes that 
a friction clutch employing a diaphragm spring represented by the curve 
440 and resilient segments represented by the curve 442 can remain in 
disengaged condition without the need for the application of any 
disengaging force to the tips of the prongs (actuating means) of the 
diaphragm spring. 
FIGS. 20, 20a, 21 and 22 illustrate a portion of a torque transmitting 
friction clutch 501, wherein the diaphragm spring 513 which performs the 
function of a sensor is coupled to the housing or cover 502 by a bayonet 
mount 514 so that the sensor 513 is maintained in a predetermined axial 
position relative to the bottom wall 502a of the cover 502. The main 
portion 513b of the sensor 513 is provided with radially outwardly 
extending coupling portions or arms 513d which are offset relative to the 
general plane of the main portion 513b in a direction toward the bottom 
wall 502a and into female coupling portions 502a' provided in the 
substantially axially extending marginal portion 502b of the cover 502; 
the marginal portion 502b surrounds the bottom wall 502a and extends 
toward the pressure plate 503 of the friction clutch 501. The female 
coupling portions 502a' which are shown in the drawing constitute lugs of 
one piece with the cover 502 and obtained as a result of appropriate 
deformation of corresponding parts of the marginal portion 502b. Each 
female coupling portion 502a' (these coupling portions form part of the 
bayonet mount 514 and are of one piece with the cover 502) is preferably 
flanked by at least one slit or slot (such as the slits 502c, 502 d) in 
the adjacent portion of the cover 502. By actually separating certain 
parts of the coupling portions 502a' from the adjacent portions of the 
cover 502, the portions 502a' can be more readily shaped to assume an 
optimum configuration for cooperation with the male coupling portions 513d 
of the bayonet mount 514. 
As can be readily seen in FIG. 21, the positions and shapes of the coupling 
portions 502a' and 513d (which together constitute the bayonet mount 514) 
are selected in such a way that they can further perform the function of 
means for centering the sensor 513 relative to the cover 502. To this end, 
the female coupling portions 502a' are provided with rather shallow 
centering recesses 502e for parts of the respective male coupling portions 
513d. 
In order to ensure predictable and optimal positioning of the sensor 513 
relative to the cover 502 during establishment of engagement between the 
coupling portions 502a' and 513d of the bayonet mount 514, the 
substantially axially extending marginal portion 502b of the cover 502 is 
preferably provided with at least three equidistant female coupling 
portions 502a'. The arrangement is such that the portions 502a' and 513d 
of the bayonet mount 514 permit a predetermined angular displacement of 
the cover 502 and the sensor 513 relative to each other before the bayonet 
mount is effective to maintain the sensor in an optimum position at a 
certain distance from the outer side of the bottom wall 502a as well as in 
properly centered position relative to the cover 502. At such time, the 
male coupling portions 513d abut stops 502f which form part of the cover 
502 and serve to prevent further rotation of the cover 502 and sensor 513 
relative to each other in order to activate the bayonet mount 514. As can 
be seen in FIG. 20a, each stop 502f can constitute an axially extending 
projection of the cover 502. FIG. 20a further shows that at least one of 
the female coupling portions 502a' (but preferably at least two or all 
three coupling portions 502a') is provided with an additional stop 502g 
which also prevents rotation of the sensor 5013 relative to the cover 502. 
Each stop 502g is engaged by the adjacent male coupling portion 513d of 
the sensor 513 when the bayonet mount 514 is fully assembled. 
In the embodiment which is shown in FIGS. 20 to 22, each of the female 
coupling portions 502a' is provided with a first stop 502f and with a 
second stop 502g for the respective male coupling portion 513d. One of the 
stops 502f, 502g holds the respective coupling portion 513d against 
rotation in one direction and the other of the stops 502f, 502g holds the 
respective coupling portion 513d against rotation in the opposite 
direction. The stops 502g serve to prevent accidental or unintentional 
separation of the bayonet mount 514, i.e., they prevent accidental 
separation of the sensor 513 from the cover 502. Once the bayonet mount 
514 is active, the sensor 513 is held in a predetermined angular position 
relative to the bottom wall 502a of the cover 502. 
In order to render the bayonet mount 514 effective, the sensor 513 is first 
subjected to an initial stress by deforming it axially toward the bottom 
wall 502a of the cover 502 so that the male coupling portions 513d can 
enter the adjacent slots or slits 502c and 502d of the cover 502 by moving 
in the circumferential direction of the friction clutch 501. In this 
manner, the male coupling portions 513d can be moved behind the adjacent 
female coupling portions 502a'. The next step of rendering the bayonet 
mount 514 operative involves turning of the cover 502 and the sensor 513 
relative to each other until at least some of the coupling portions 513d 
reach and are arrested by the corresponding stops 502f. The sensor 513 
then dissipates some energy so that at least some of the male coupling 
portions 513d move axially and away from the bottom wall 502a and enter 
the spaces between the respective stops 502f and 502g. This ensures that 
the sensor 513 can no longer become accidentally separated from the 
(female) coupling portions 502a' of the cover 502. Once the bayonet mount 
514 is effective to reliably hold the sensor 513 in the interior of the 
cover 502, the assembly of the friction clutch 501 can proceed without 
risking accidental changes in the (centered) position of the sensor 513 
relative to the cover 502 and/or unintentional separation of the sensor 
from the cover. At such time, each of the female coupling portions 502a' 
is overlapped by one of the male coupling portions 513d. 
In the heretofore described embodiments of the improved friction clutch, 
that circumferentially complete portion of the sensor (such as the main 
portion 513b of the sensor 513) which actually generates the force (e.g., 
the force to urge the diaphragm spring 504 of FIG. 21 against the ring 
512) extends radially outwardly beyond the points or lines of contact 
between the diaphragm spring and the pressure plate (such as the diaphragm 
spring 504 and the portion 503a of the pressure plate 503 shown in FIG. 
21). However, it is often desirable and advantageous to position the main 
portion of the sensor radially inwardly of the locations of engagement 
between the diaphragm spring and the pressure plate, i.e., radially 
inwardly of the circle including the points or lines of contact between 
the diaphragm spring and the pressure plate. With reference to the 
friction clutch 1 of FIGS. 1 and 2, this would mean that the 
circumferentially complete portion 13b of the sensor 13 would be located 
radially inwardly of the points of contact between the diaphragm spring 4 
and the projecting portions 3a of the pressure plate 3. 
Referring again to the friction clutch 501 of FIGS. 20 to 22, the ramps 524 
of the adjusting unit 516 are provided directly at the inner side of the 
bottom wall 502a of the cover 502. The latter is made of sheet metal and 
the ramps 524 are obtained by appropriate deformation of an annular 
portion of the bottom wall 502a. The means for biasing the annular member 
517 of the adjusting unit 516 includes coil springs 526 which are guided 
by suitably curved guide elements or mandrels 528 forming part of the 
member 517 (see particularly FIG. 22). The coil springs 526 react against 
the cover 502 and bear against the respective projections 527 of the 
member 517 so that the latter tends to turn in a direction to move (under 
the action of the ramps 524) toward the pressure plate 503 and to thus 
compensate for wear upon the pressure plate 503, the counterpressure plate 
(not shown in FIGS. 20-22) and/or friction linings 507 between the 
counterpressure plate and the pressure plate 503. As can be seen in FIG. 
21, each mandrel 528 can have an elongated rectangular cross-sectional 
outline to extend substantially diametrically across the entire space 
within the surrounding convolutions of the respective coil spring 526. The 
length of the arcuate mandrels 528 can approximate but can be less than 
the length of the respective coil springs 526. The utilization of 
relatively long mandrels 528 ensures predictable and satisfactory guidance 
of the respective coil springs 526 at least in the radial direction of the 
member 517. In addition, the mandrels 528 can be designed and dimensioned 
to effectively prevent any, or any appreciable, axial movements (buckling) 
of intermediate portions of the respective coil springs 526. Another 
important advantage of the mandrels 528 is that they simplify the assembly 
of the friction clutch 501. 
FIG. 22 shows one of several radially inwardly extending projections 527 
which are or can be of one piece with the major portion of the member 517 
and carry the respective mandrels 528. If the member 517 is made of a 
plastic material (e.g., a material which can be shaped in an injection 
molding or extruding machine), the projections 527 can be made of one 
piece with the respective mandrels 528 as well as with the 
circumferentially complete main portion of the member 517, namely that 
portion which is provided with ramps 518 serving to cooperate with the 
ramps 524 on the bottom wall 502a of the cover 502. However, it is equally 
within the purview of the invention to mass produce the mandrels 528 (or 
the mandrels 528 and the corresponding projections 527) independently of 
the main portion of the member 517 and to thereupon assemble the parts 527 
or the parts 527, 528 with the main portion of the member 517, e.g., by 
resorting to connections which operate with snap action. It is also 
possible to make the mandrels 528 from a one-piece ring which is severed 
at a required number of locations to permit entry of the thus obtained 
arcuate portions of the ring into the corresponding coil springs 526 and 
to affix each arcuate portion of the subdivided ring to one of the 
projections 527. The connections between the arcuate portions of the 
aforementioned ring (i.e., of a blank for the making of the mandrels 528 
or their equivalents) and the projections 527 can be designed to operate 
by snap action. It is preferred to provide the member 517 with at least 
three preferably equidistant projections 527. 
If desired or necessary, the friction clutch 501 can be constructed in such 
a way that it comprises one or more additional systems for preventing 
undesirable movements of the coil springs 526 relative to the cover 502 
and/or member 517. For example, and as already explained with reference to 
the friction clutch 101 of FIGS. 12-13, the cover 502 and/or the member 
517 can be provided with suitable means for preventing any undesirable 
movements of the coil springs 526 under the action of centrifugal force. 
The means for coupling one end of each coil spring 526 to the cover 502 of 
the friction clutch 501 comprises retainers or stops 526a' (one shown in 
each of FIGS. 20 and 21) which can constitute suitably deformed portions 
of the cover 502 and extend in the axial direction of the friction clutch. 
The configuration of the retainers 526a' is preferably such that they not 
only abut the adjacent outermost convolutions of the respective coil 
springs 526 but that they are also capable of otherwise guiding or 
locating the respective coil springs (e.g., in the radial and/or axial 
direction of the friction clutch 501). 
In the friction clutch 601 of FIG. 23, the sensor 613 is located at the 
outer side of the bottom wall 602a of the housing or cover 602, i.e., at 
that side of the bottom wall 602a which faces away from the pressure plate 
603. An advantage of such mounting of the sensor 613 is that it is 
subjected to less pronounced thermal stresses; this reduces the likelihood 
of undesirable reduction or decrease of resiliency of the sensor 613 as a 
result of excessive thermal stressing. Moreover, the sensor 613 at the 
outer side of the bottom wall 602a is subjected to much more pronounced 
cooling action when the friction clutch 601 is in use. 
The operative connection between the sensor 613 and the diaphragm spring 
604 in the clutch 601 of FIG. 23 is established by way of distancing 
elements in the form of rivets 615 (only one shown). The shanks of these 
rivets extend through slots between the neighboring prongs of the 
diaphragm spring 604 and through openings in the bottom wall 602a of the 
cover 602. The axes of the rivets 615 are parallel to the axis of the 
friction clutch 601, and each of these rivets has a head which overlies 
the outer side of the sensor 613. The rivets 615 constitute but one form 
of means which can be used to operatively connect the sensor 613 with the 
diaphragm spring 604. For example, the sensor 613 can be provided with 
axially extending projections in the form of lugs or the like having 
suitable tips overlying the ring 611 of the seat 605 to maintain the ring 
611 in uninterrupted contact with the main portion of the diaphragm spring 
604. In fact, it is possible to design the sensor 613 in such a way that 
it is made of one piece with parts which replace the rivets 615 as well as 
the ring 611 of the seat 605. 
Referring to FIG. 24, there is shown a portion of a friction clutch 701 
with a sensor 713 which is located radially inwardly of the locations of 
contact between the diaphragm spring 704 and the portion or portions 703a 
of the pressure plate 703. Thus, the sensor 713 is located radially 
inwardly of the seat 705. The radially inner portions (tongues) of the 
sensor 713 react against the adjacent portions of the cover 702. To this 
end, the cover 702 is provided with arms 715 which extend through the 
slots between the prongs of the diaphragm spring 704 and are engaged by 
the adjacent portions of the sensor 713. It is equally possible to provide 
the radially inner portion of the sensor 713 with arms which extend 
through slots between the prongs of the diaphragm spring 704 and engage 
the cover 702. Instead of extending through the slots between the prongs 
of the diaphragm spring 704, the aforementioned arms of the sensor 713 can 
extend through specially provided openings in the diaphragm spring 704. 
The annular adjusting member 817 which is shown in FIG. 25 can be utilized 
with advantage in the friction clutch of FIGS. 20, 20a and 21 in lieu of 
the annular member 517 of FIG. 22. The radially inner portion of the 
member 817 is provided with projections 827 which extend radially inwardly 
and have radially inwardly projecting extensions 827a. The extensions 827a 
serve as abutments for the adjacent end convolutions of arcuate coil 
springs 826 extending in the circumferential direction of the member 817. 
The other end convolution of each coil spring 826 bears against a retainer 
826a forming part of a housing or cover (not shown but corresponding to 
the cover 502 of FIGS. 20 and 21) and extending in parallelism with the 
axis of the friction clutch employing the member 817. 
In order to facilitate assembly of the member 817 with the coil springs 
826, there is provided a split ring 828 which is concentric or nearly 
concentric with the member 817 and extends through the extensions 827a, 
through the coil springs 826 and through the retainers 826a. The ring 828 
is affixed to the extensions 827a; for example, the extensions 827a can be 
provided with grooves or sockets extending in the circumferential 
direction of the member 817 and being dimensioned and configurated to 
receive the respective portions of the split ring 828 by snap action. Each 
retainer 826a can be provided with a groove 826b which extends in 
substantial parallelism with the axis of the member 817 and is 
configurated and dimensioned to receive the adjacent portion of the split 
ring 828 with freedom of movement of the ring relative to the retainer 
826a in the circumferential direction of the member 817. At the very 
least, the ring 828 can move circumferentially of the member 817 to the 
extent which is necessary to compensate for wear upon the friction 
linings, the pressure plate and/or the counterpressure plate in the 
friction clutch which employs the structure of FIG. 25. 
It is presently preferred to configurate the projections 827 and the 
retainers 826a in such a way that the sockets of the extensions 827a (for 
reception of the adjacent portions of the split ring 828 by snap action) 
face in one axial direction and the grooves 826b (for reception of 
adjacent portions of the ring 828 with freedom of movement in the 
circumferential direction of the member 817) face in the opposite axial 
direction. In other words, the sockets of the extensions 827a can be open 
in a direction toward or away from the bottom wall of the housing or cover 
of the friction clutch employing the structure of FIG. 25, and the grooves 
826b of the retainers 826a can be open in a direction away from the bottom 
wall of such housing or cover. 
FIG. 26 illustrates a friction clutch 901 with a diaphragm spring 904 
having a main portion 904a. The median part of the main portion 904a is in 
contact with the parts of the seat 905 and the radially outermost part of 
the main portion 904a is positioned to bear against the projecting portion 
or portions 903a of the pressure plate 903 when the friction clutch 901 is 
engaged. The prongs 904b of the diaphragm spring 904 (i.e., the actuating 
means of the means for engaging and disengaging the friction clutch 901) 
extend radially inwardly beyond the main portion 904a, i.e., radially 
inwardly beyond the seat 905. The distance of the seat 905 from the 
radially innermost part of main portion 904a of the diaphragm spring 904 
is greater than in heretofore known friction clutches wherein the means 
for biasing the pressure plate toward the friction linings of the clutch 
disc includes a diaphragm spring reacting against the housing or cover of 
the friction clutch. In the embodiment of FIG. 26, the ratio of the width 
of that part of the main portion 904a which extends radially inwardly 
beyond the seat 905 to the width of that part of the main portion 904a 
which extends radially outwardly beyond the seat 905 is approximately 1:2. 
It is often desirable that such ratio be between 1:6 and 1:2. By selecting 
the position of the seat 905 relative to the main portion 904a of the 
diaphragm spring 904 in the just outlined manner, the maker of the 
friction clutch 901 reduces the likelihood of damage to and/or 
overstressing of the main portion 904a in the region of engagement with 
the seat 905. In other respects, the friction clutch 901 of FIG. 26 can be 
constructed and assembled in a manner as described with reference to the 
friction clutch 101 of FIGS. 12 and 13. 
FIG. 26 further shows, by broken lines, an axially extending centering 
projection 903b on the illustrated axially projecting portion 903a of the 
pressure plate 903. The pressure plate 903 can be provided with a 
circumferentially complete projecting portion 903a or with a discontinuous 
projecting portion, e.g., with at least three equidistant discrete 
projecting portions 903a. The single projecting portion or each discrete 
projecting portion 903a of the pressure plate 903 can be provided with a 
centering projection 903b for the diaphragm spring 904. The centering 
projections 903b render it possible to dispense with all other means for 
centering the diaphragm spring 904 relative to the bottom wall 902a of the 
housing or cover 902. Though FIG. 26 shows a rivet 915 which is to center 
the diaphragm spring 904, such rivet is optional if the projecting portion 
or portions 903a of the pressure plate 903 are provided with centering 
projections 903b. 
It is further possible to replace the rivets 915 and/or the centering 
projection or projections 903b of the pressure plate 903 in the friction 
clutch 901 of FIG. 26 with a set of centering projections which are of one 
piece with or are affixed (e.g., welded) to the bottom wall 902a of the 
cover 902. For example, the centering projections of the cover 902 can 
constitute lugs which are bent out of the bottom wall 902a and extend in 
parallelism with the axis of the friction clutch 901 toward the pressure 
plate 903. Alternatively, the centering projections of the cover 902 can 
constitute inwardly bulging portions (rather than lugs) of the bottom wall 
902a. 
The diaphragm spring which constitutes the sensor 913 in the friction 
clutch 901 of FIG. 27 is designed in such a way that its circumferentially 
complete main or basic portion 913a is disposed radially inwardly of the 
projecting portion or portions 903a of the pressure plate 903. In order to 
prop the diaphragm spring 904 on the one hand, and to be adequately 
propped against the cover 902 on the other hand, the sensor 913 is further 
provided with radial arms in the form of tongues including a set extending 
from the main portion 913a radially inwardly to form part of the seat 905 
(such as a substitute for the wire ring 11 in the seat 5 of FIGS. 1-2) and 
a set extending from the main portion 913a radially outwardly to react 
against lugs forming part of the substantially axially extending portion 
of the cover 902. 
Referring to FIG. 27, there is shown a friction clutch 1001 including a 
diaphragm spring 1013 constituting a sensor and serving to transmit a 
force which opposes the force to be applied in order to disengage the 
friction clutch and which also opposes the force of the diaphragm spring 
(resilient device) 1004. The sensor 1013 reacts against the housing or 
cover 1002 and bears against the projecting portion or portions 1003a of 
the pressure plate 1003. In other words, the sensor 1013 is installed in 
axially stressed condition between the cover 1002 and the pressure plate 
1003. In this embodiment of the present invention, the seat 1005 does not 
provide for the diaphragm spring 1004 a bearing for tilting of the 
diaphragm spring in a direction to disengage the friction clutch 1001. The 
diaphragm spring 1004 engages a wire ring 1012 which forms part of the 
seat 1005 and contacts that side of the main portion of the diaphragm 
spring 1004 that faces toward the annular adjusting member 1017 and the 
bottom wall 1002a of the cover 1002. The sensor 1013 constitutes the means 
for biasing the main portion of the diaphragm spring 1004 against the wire 
ring 1012 of the seat 1005. The sensor 1013 is dimensioned, configurated 
and installed in such a way that, during disengagement of the friction 
clutch 1001, the axial force generated by the sensor 1013 and acting upon 
the diaphragm spring 1004 is or becomes larger than the force which is to 
be applied to disengage the friction clutch 1001. The arrangement should 
be such as to ensure that, when the wear upon the friction linings (not 
shown in FIG. 27) is nil or minimal, the diaphragm spring 1004 
continuously engages the wire ring 1012 of the seat 1005. To this end, and 
as already described in connection with the previously discussed 
embodiments of the improved friction clutch, it is necessary to properly 
relate the superimposed forces acting in the axial direction of the 
friction clutch 1001. Such forces are generated by the sensor 1013, by the 
resilient segments (not shown) of the clutch disc in the friction clutch 
1001, by leaf springs (if any) which connect the pressure plate 1003 with 
the cover 1002 in such a way that the parts 1002, 1003 have a certain 
freedom of axial movement but cannot turn relative to each other, by the 
diaphragm spring 1004, by the means for disengaging the friction clutch 
1001, and by resilient means (e.g., coil springs or torsion springs) 
acting upon the member 1017 of the adjusting unit in order to compensate 
for wear upon the pressure plate 1003, the counterpressure plate (not 
shown) and/or the friction linings of the clutch disc between the pressure 
plate 1003 and the counterpressure plate. 
FIGS. 28 to 32 illustrate a further torque transmitting friction clutch 
1101 having a housing or cover 1102 and a pressure plate 1103. The latter 
is connected with the cover 1102 in the aforedescribed manner, i.e., with 
some freedom of axial movement but without any freedom of angular 
movement. A diaphragm spring 1104 is installed in the cover 1002 between 
the bottom wall 1002a and the pressure plate 1103 to bias the pressure 
plate 1103 against the adjacent set of friction linings 1107 forming part 
of a clutch plate or clutch disc 1108. When the diaphragm spring 1104 is 
free to bias the pressure plate 1103 in a direction away from the bottom 
wall 1102a of the cover 1102, the other set of friction linings 1107 of 
the clutch disc 1108 is caused to bear against the friction surface of a 
counterpressure plate 1106, e.g., a flywheel or a portion of a flywheel 
which receives torque from a suitable prime mover, particularly from the 
output element (e.g., a crankshaft) of an internal combustion engine in a 
motor vehicle. The clutch disc 1108 then transmits torque to the input 
element of a variable-speed transmission in the power train between the 
flywheel 1106 and the wheels of the motor vehicle. 
The means for non-rotatably but axially movably connecting the pressure 
plate 1103 to the cover 1102 comprises a set of substantially tangentially 
extending leaf springs 1109 (FIG. 28). 
The clutch disc 1108 comprises resilient segments 1110 which are designed 
and mounted to establish a progressive buildup of torque during engagement 
of the friction clutch 1101. This is achieved in that the segments 1110 
permit the two sets of friction linings 1107 (namely the set engageable by 
the friction surface of the pressure plate 1103 and the set engageable by 
the friction surface of the flywheel 1106) to perform limited axial 
movements toward each other and to thus permit a progressive buildup of 
forces acting upon the friction linings 1107 in the axial direction of the 
clutch 1101. However, it is equally within the purview of the invention to 
employ a modified clutch disc wherein the two sets of friction linings are 
installed at a fixed axial distance from each other, e.g., by being bonded 
or otherwise affixed to opposite sides of at least one rigid washer-like 
carrier extending radially outwardly from the hub 1108a of the clutch 
disc. In the thus modified friction clutch 1101, a functional equivalent 
of the resilient segments 1110 can be installed at another point, e.g., 
between the diaphragm spring 1104 and the pressure plate 1103 and/or 
between the cover 1102 on the one hand and the pressure plate 1103 or the 
flywheel 1106 on the other hand. 
In the embodiment of FIGS. 28 to 32, the diaphragm spring 1104 comprises a 
circumferentially complete main portion 1104a and prongs 1104b (actuating 
means of the means for engaging and disengaging the friction clutch 1101) 
which are of one piece with and extend radially inwardly from the main 
portion 1104a. The radially outer part of the main portion 1104a biases 
the pressure plate 1103 against the adjacent friction linings 1107 when 
the friction clutch 1101 is engaged, and a radially inner part of the main 
portion 1104a cooperates with the seat 1105 in order to ensure that the 
diaphragm spring can be tilted relative to the bottom wall 1102a of the 
cover 1102. The seat 1105 includes two annular portions 1111 and 1112 
which are disposed at opposite sides of the main portion 1104a of the 
diaphragm spring 1104 and each of which can constitute a wire ring. These 
wire rings provide a bearing which enables the corresponding part of the 
main portion 1104a of the diaphragm spring 1104 to be tilted in order to 
urge the pressure plate 1103 toward the flywheel 1106 or to permit the 
pressure plate to move away from the flywheel, e.g., under the bias of the 
leaf springs 1109. 
The means for preventing rotation of the diaphragm spring 1104 relative to 
the bottom wall 1102a and for centering the diaphragm spring in the cover 
1102 comprises a set of distancing elements in the form of rivets 1115 
which are anchored in the bottom wall 1102a and extend in parallelism with 
the axis of the friction clutch 1101. The shanks 1115a of the rivets 1115 
extend through slots between the neighboring prongs 1104b of the diaphragm 
spring 1104. 
The friction clutch 1101 further comprises means for compensating for wear 
upon the pressure plate 1103, upon the flywheel 1106 and particularly upon 
the friction linings 1107 of the clutch disc 1108. Such compensating means 
comprises an adjusting unit 1116 which operates between the pressure plate 
1103 and the diaphragm spring 1104, as well as a device 1117 which limits 
the extent of movability of the pressure plate 1103 in a direction away 
from the flywheel 1106, i.e., in a direction to release the clutch disc 
1108 and to thus disengage the friction clutch 1101. The device 1117 can 
be said to constitute a means for monitoring the extent of axial movements 
of the pressure plate 1103 relative to the flywheel 1106 and/or the bottom 
wall 1102a of the cover 1102. 
The monitoring device 1117 ascertains the extent of wear upon the pressure 
plate 1103, the flywheel 1106 and/or the friction linings 1107 and 
comprises a set of sleeves 1118 each of which is non-rotatably installed 
in a bore or hole 1120 of the pressure plate 1103. Each sleeve 1118 has an 
axially parallel slot or passage 1121 to permit entry of a pair of 
resilient elements 1122 in the form of leaf springs in the axial direction 
of the friction clutch 1101. The leaf springs 1122 of each pair bear 
against each other and at least one leaf spring 1122 of each pair can have 
an arcuate shape. It is presently preferred to utilize pairs of leaf 
springs 1122 wherein each leaf spring has an arcuate shape, and the leaf 
springs of each pair are bent in opposite directions, for example, in such 
a way that a convex side of one leaf spring 1122 of each pair bears 
against a convex side of the other leaf spring 1122 of the respective 
pair. 
The leaf springs 1122 of each pair are installed in the respective sleeves 
1118 with a predetermined amount of initial stress so that a certain 
predetermined frictional resistance must be overcome before the leaf 
springs of a pair of such springs can move relative to the respective 
sleeve 1118 in the axial direction of the friction clutch 1101. The length 
of the leaf springs 1122 in the axial direction of the friction clutch 
1101 is selected in such a way that, when the friction clutch is engaged 
so that the friction linings 1107 are clamped between the neighboring 
friction surfaces of the pressure plate 1103 and flywheel 1106, a certain 
clearance or gap 1124 is maintained between the leaf springs and an 
axially fixed part of the friction clutch, e.g., the marginal portion 1123 
of the cover 1102. The width of the clearance 1124 corresponds to the 
predetermined axial movability of the pressure plate 1103 relative to the 
cover 1102. When the friction clutch 1101 is engaged, those ends 1122a of 
the leaf springs 1122 which are remote from the cover 1102 come into 
abutment with the flywheel 1106; this ensures that the pressure plate 1103 
is moved axially of the friction clutch 1101 with reference to the leaf 
springs 1122 to an extent which exactly corresponds to the extent of wear 
upon the pressure plate 1103, upon the flywheel 1106 but mainly or 
exclusively upon the friction linings 1107. The displacement of the 
pressure plate 1103 relative to the leaf springs 1122 takes place against 
the opposition of the friction existing between the leaf springs 1122 and 
the respective sleeves 1118. These sleeves can be made of a plastic 
material or from another material which preferably has a high coefficient 
of friction. 
In the illustrated embodiment, the bores or holes 1120 receive the sleeves 
1118 in such a way that each sleeve is a press fit therein and is held 
against movement in the axial as well as in the circumferential direction. 
The bores or holes 1120 are provided in lobes 1125 of the pressure plate 
1103. Each lobe 1125 (only one shown in FIG. 28) extends radially 
outwardly and further serves to carry one end portion of one of the leaf 
springs 1109. The connections between the leaf springs 1109 and the 
respective lobes 1125 of the pressure plate 1103 include rivets 1109a. A 
shifting of sleeves 1118 in a direction toward the flywheel 1106 can also 
be avoided or prevented in that each sleeve 1118 is provided with a collar 
1118a at that end which is adjacent the cover 1102. Such collar engages 
the pressure plate 1103. Any movement of a sleeve 1118 in the respective 
bore or hole 1120 in a direction toward the cover 1102 can be prevented by 
such configuration of the leaf springs 1109 (as shown in FIG. 28 by broken 
lines at 1119) that each leaf spring 1109 partially overlies the 
respective sleeve 1118 and, if necessary, fixedly secures the sleeve 1118 
in its bore or hole 1120. Turning of a sleeve 1118 in its bore or hole 
1120 can also be prevented by imparting to each sleeve a profile (e.g., by 
providing it with an extension) which receives the portion 1119 of the 
respective leaf spring 1109. 
The adjusting unit 1116 comprises a compensating element in the form of a 
ring 1126 having a U-shaped cross-sectional outline and being acted upon 
by the diaphragm spring 1104. A substantial portion of the ring 1126 is 
shown in FIG. 32 and a portion of this ring (as seen from its open side) 
is also shown in FIG. 30. The ring 1126 comprises a circumferentially 
complete bottom wall or end wall 1127, a circumferentially complete 
radially inner cylindrical sidewall 1130 and a circumferentially complete 
radially outer cylindrical sidewall 1131. That side of the end wall or 
bottom wall 1127 which faces the diaphragm spring 1104 is provided with at 
least one ring-shaped axial projection 1128. If the wall 1127 carries 
several projections 1128, they are preferably equidistant from each other 
in the circumferential direction of the ring 1126. Furthermore, if the 
ring 1126 is made of a metallic sheet material, the projections 1128 can 
constitute deformed portions of the end wall 1127. If the projections 1128 
are segment shaped, the neighboring projections 1128 define radial 
passages between the main portion 1104a of the diaphragm spring 1104 and 
the ring 1126; such passages permit circulation of air to achieve 
desirable cooling of the corresponding portion of the friction clutch 
1101. 
Referring to FIG. 29, the ring 1126 is centered relative to the pressure 
plate 1103 by at least one shoulder 1129 which is provided on the pressure 
plate 1103 adjacent the inner sidewall 1130 of the ring 1126. The shoulder 
1129 can constitute a circumferentially complete surface of the pressure 
plate 1103 or a composite surface consisting of a plurality of discrete 
arcuate sections adjacent the outer side of the sidewall 1130 of the ring 
1126. 
The walls 1127, 1130, 1131 of the ring 1126 define a ring-shaped space 
1126a and the outer wall 1131 is provided with a set of equidistant 
projections or lobes 1132 which extend radially outwardly and cooperate 
with complementary projections 1133 of the axially movable leaf springs 
1122 forming part of the monitoring device 1117. The complementary 
projections 1133 can constitute suitably shaped integral parts of the leaf 
springs 1122 and extend radially inwardly to overlie and to be thus 
located in the path of movement of projections 1132 on the radially outer 
sidewall 1131 of the ring 1126. This ensures that the ring 1126 cannot 
move away from the pressure plate 1103 in a direction toward the cover 
1102. 
A displacing device 1134 between the ring 1126 and the pressure plate 1103 
serves to automatically reset the ring 1126 during disengagement of the 
friction clutch 1101 in order to compensate for wear upon the pressure 
plate 1103 and/or flywheel 1106 but mainly for wear upon the friction 
linings 1107. The displacing device 1134 performs a self-locking 
(blocking) action during engagement of the friction clutch 1101 to thus 
ensure that the ring 1126 will assume a predetermined axial position 
relative to the pressure plate 1103 while the friction clutch is in the 
process of being engaged. The position of the ring 1126 relative to the 
pressure plate 1103 can change only during disengagement of the friction 
clutch 1101 and only to the extent determined by the amount of wear upon 
the aforementioned parts 1103, 1106 and/or 1107. 
The displacing device 1134 comprises a plurality of pairs of wedges 1135, 
1136, and such pairs are preferably equidistant from each other in the 
circumferential direction of the ring 1126. The pairs of wedges 1135, 1136 
are installed in the circular internal space 1126a of the ring 1126. The 
wedges 1136 contact a ring-shaped surface 1137 of the pressure plate 1103 
and are non-rotatably secured to the ring 1126 but are axially movably 
installed in the internal space 1126a. To this end, the sidewalls 1130, 
1131 of the ring 1126 are provided with guide means in the form of ribs 
1138, 1139 constituting projections extending into the space 1126a and 
confining the wedges 1136 to movement in the axial direction of the 
friction clutch 1101. To this end, the wedges 1136 are provided with 
recesses or grooves 1140, 1141 to receive the ribs 1138, 1139, 
respectively. The grooves 1140, 1141, as well as the ribs 1138, 1139, 
extend in parallelism with the axis X--X of the friction clutch 1101. 
As can be seen in FIG. 30, the wedges 1135 are installed axially in the 
space 1126a of the ring 1126, namely between the bottom wall or end wall 
1127 and the adjacent wedges 1136. The wedges 1135, 1136 respectively 
include or define ramps 1143, 1142 which extend in the circumferential 
direction of the ring 1126 and slope axially of the friction clutch 1101 
in a direction from the inner side of the end wall 1127 toward the open 
side of the ring 1126. The ramps 1143 of the wedges 1135 abut the ramps 
1142 of the adjacent wedges 1136. Those sides of the wedges 1135 which 
face away from the respective ramps 1143 are adjacent the inner side of 
the end wall 1127, and each wedge 1135 can be shifted in the 
circumferential direction of the ring 1126 in order to compensate for wear 
upon the pressure plate 1103, flywheel 1106 and/or friction linings 1107. 
The ramps 1142 of the wedges 1136 bear against the ramps 1143 of the 
adjacent wedges 1135. Such engagement between the ramps 1142 and the 
adjacent ramps 1143 is achieved by the provision of coil springs 1144 
which are received in the space 1126a of the ring 1126. Each spring 1144 
reacts against one of the wedges 1136 (which are held against movement in 
the circumferential direction of the ring 1126) and bears against one of 
the wedges 1135 (i.e., against one of those wedges which are movable in 
the circumferential direction of the ring 1126). The wedges 1135, 1136 are 
respectively provided with projections 1145, 1146 which extend into the 
adjacent end convolutions of the respective springs 1144. These springs 
are further confined and guided by the internal surfaces of the walls 
1127, 1130 and 1131 of the ring 1126. 
The friction clutch 1101 is constructed in such a way that the ring 1126 
cannot rotate relative to the pressure plate 1103. As can be seen in FIG. 
31, the pressure plate 1103 is provided with axially extending projections 
in the form of pins or studs 1147 extending through holes or bores 1148 
provided in the aforementioned projections 1132 extending radially 
outwardly from the sidewall 1131 of the ring 1126. Such non-rotatable 
mounting of the ring 1126 relative to the pressure plate 1103 ensures 
that, when the friction clutch 1101 is in use, the projections 1132 are 
always overlapped by the projections 1133 of the leaf springs 1122. 
The wedges 1135, 1136 in the space 1126a of the ring 1126 are assumed to be 
made of a heat-resistant plastic material, such as a thermoplastic or a 
pressure setting substance. The material of these wedges can be reinforced 
by filaments of glass fibers or the like. Such construction of the wedges 
1135, 1136 is preferred at this time because it renders it possible to 
mass produce the wedges in an injection molding or other readily available 
plastic processing machine. However, it is also possible, and often 
preferred, to make at least one set of the wedges 1135, 1136 of a material 
having a high coefficient of friction, e.g., from the material of which 
the friction linings 1107 are made. Still further, it is possible to make 
the wedges 1135 and/or 1136 of metallic sheet material or from a suitable 
sintered metal. 
The inclination and the length of the ramps 1142, 1143 are selected in such 
a way that one ensures reliable adjustment of the wedges 1135, 1136 
relative to each other in order to compensate for wear upon the pressure 
plate 1103, flywheel 1106 and/or friction linings 1107 during the entire 
useful life of the friction clutch 1101. The inclination (angle 1149 in 
FIG. 30) of the ramps 1142, 1143 with reference to a plane which is normal 
to the axis X--X of the friction clutch 1101 is selected with a view to 
ensure that friction which develops when the ramps 1142, 1143 are biased 
against each other suffices to prevent any slippage of the wedges 1135 and 
the associated wedges 1136 relative to each other. The magnitude of the 
angles 1149 (slope of the ramps 1142, 1143) will depend upon the selection 
of materials of the wedges 1135, 1136 and is normally between 
approximately 5 and 20 degrees, preferably close to or exactly 10 degrees. 
The wedges 1135 which can move in the circumferential direction of the 
ring 1126 are oriented in such a way that their tips face in the direction 
(arrow 1150) of rotation of the friction clutch 1101. Furthermore, the 
magnitude of the angles 1149 and the bias of the springs 1144 are such 
that the resultant axial force acting upon the ring 1126 is smaller than 
the force which is required to move the leaf springs 1122 of the 
monitoring device 1117. 
It is further important or desirable to select the characteristics of the 
diaphragm spring 1104 in such a way that the force to be applied by this 
spring against the pressure plate 1103 can be increased by a value 
corresponding to the force which is needed to displace the leaf springs 
1122 plus the stressing of leaf springs 1109 between the cover 1102 and 
the pressure plate 1103. Furthermore, the component parts of the friction 
clutch 1101 should be designed in such a way that, in comparison with the 
wear upon the friction linings 1107, the wear at the locus or loci of 
engagement of the ring 1126 with the diaphragm spring 1104 as well as the 
wear between the leaf springs 1122 and the flywheel 1106 and between the 
leaf springs 1122 and the cover 1102 be small or negligible. 
Referring again to FIG. 30, and in order to avoid unintentional shifting of 
the ramps 1142, 1143 of the wedges 1136, 1135 relative to each other, at 
least one of each pair of cooperating ramps 1142, 1143 can be provided 
with relatively small projections or protuberances which tend to be caught 
by the adjacent ramps. These protuberances can be designed and dimensioned 
with a view to permit necessary movements of the pairs of wedges 1135, 
1136 relative to each other in order to compensate for wear upon the parts 
1103, 1106 and/or 1107 but to prevent any undesirable slippage of the 
ramps 1142, 1143 relative to one another. It is normally preferred to 
provide the just discussed minute protuberances on each of the ramps 1142 
as well as on each of the ramps 1143 and to orient the protuberances in 
such a way that those on the ramps 1142 mate or mesh or become interlaced 
with the protuberances of the ramps 1143. For example, the protuberances 
can constitute relatively small (e.g., minute) sawtooth-shaped profiles on 
the ramps 1142 and/or 1143. Such protuberances should be capable of 
preventing accidental or unintentional shifting of the pairs of wedges 
1135, 1136 relative to each other but they should not interfere with those 
adjustments of the ramps 1142, 1143 relative to each other which are 
needed to compensate for the aforediscussed wear, mainly upon the friction 
linings 1107 but preferably also (if any) upon their friction surfaces of 
the pressure plate 1103 and flywheel 1106. FIG. 30 shows (enlarged for 
better illustration) protuberances 1143a in the form of sawteeth which are 
applied to one of the two abutting ramps 1142, 1143. If only one of the 
ramps 1142, 1143 is provided with protuberances 1143a and/or analogous 
protuberances, they can be designed in such a way that their hardness 
exceeds the hardness of the material of the adjacent (non-profiled or 
non-serrated) ramps; this ensures that the relatively hard protuberances 
will be capable of penetrating (to a small or minute extent) into the 
adjacent ramps 1142 or 1143 to thus further reduce the likelihood of 
accidental displacement of the wedges 1136 relative to the adjacent wedges 
1135 in the circumferential direction of the ring 1126. 
In the absence of any undertakings to the contrary, the temperature of the 
arcuate leaf springs 1122 would be likely to rise to a rather high value 
in response to engagement of the pressure plate 1103 with the adjacent set 
of friction linings 1107. This could result in a reduction of resiliency 
of the leaf springs 1122. Therefore, the sleeves 1118 for the leaf springs 
1122 are preferably made of a material exhibiting a low heat conductivity 
and a high friction coefficient to prevent excessive transfer of heat from 
the pressure plate 1103 during engagement of the friction clutch 1101. The 
material of the wedges 1135, 1136 can be the same as that of the sleeves 
1118. 
In order to ensure satisfactory cooling of the friction clutch 1101, 
especially of the pressure plate 1103, the latter can be provided with 
substantially radially extending grooves, channels and/or other passages 
which are preferably equidistant from each other in the circumferential 
direction of the cover 1102 and one of which is shown in FIG. 29 by broken 
lines, as at 1151. The arrangement may be such that the passages 1151 
alternate with pairs of wedges 1135, 1136 in the circumferential direction 
of the ring 1126. Each passage 1151 is provided in the pressure plate 1103 
between the adjacent set of friction linings 1107 and the ring 1126. The 
cooling action can be enhanced still further by providing the ring 1126 
with axially extending slots starting at the bottom wall or end wall 1127 
adjacent the coil springs 1144. This establishes radially extending 
passages between the diaphragm spring 1104 and the ring 1126. 
The resistance of various selected parts of the friction clutch 1101 to 
wear can be enhanced by providing such parts with coats consisting of 
suitable wear-resistant material. For example, certain parts can be 
provided with layers of hard chrome or molybdenum. Alternatively, selected 
parts of the friction clutch can be provided with inserts, shells or 
envelopes of highly wear-resistant material. For example, the leaf springs 
1122 can be provided with shoes of plastic material, at least in the 
regions where these leaf springs contact or are likely to contact the 
flywheel 1106 and/or the cover 1102. 
The leaf springs 1109 which transmit torque between the pressure plate 1103 
and the cover 1102 are installed in stressed condition in such a way that 
they shift the pressure plate in a direction toward the bottom wall 1102a 
of the cover 1102 in response to disengagement of the friction clutch 
1101. This ensures that the ring 1126 continues to abut the diaphragm 
spring 1104 during the entire stage of disengagement of the friction 
clutch 1101, i.e., until the monitoring device 1117 becomes effective. 
The path of movement of the tips 1104c of prongs 1104b of the diaphragm 
spring 1104 (i.e., of the actuating means in the means for engaging and 
disengaging the friction clutch 1101) is preferably selected in such a way 
that the radially outermost portion of the diaphragm spring 1104 is 
slightly spaced apart from the ring 1126 when the movement of the tips 
1104c in a direction to disengage the friction clutch is completed. Thus, 
when the friction clutch is being disengaged, the distance covered by the 
diaphragm spring 1104 in the region of the seat 1105 (where the diaphragm 
spring bears (directly or indirectly) against the pressure plate 1103) 
exceeds the extent of movement (clearance 1124) of the pressure plate 1103 
away from the flywheel 1106 (as determined by the leaf springs 1122). 
Those relative positions of various parts of the friction clutch 1101 which 
are shown in FIG. 29 are assumed by these parts when the extent of wear 
upon the pressure plate 1103, flywheel 1106 and friction linings 1107 is 
minimum or nil. Once the friction linings 1107 have undergone a certain 
amount of wear, the pressure plate 1103 changes its position in a 
direction toward the flywheel 1106 (when the friction clutch 1101 is 
engaged). This results in a change of conicity of the diaphragm spring 
1104 as well as in a change of bias of the diaphragm spring upon the 
pressure plate 1103, preferably in a sense to increase the bias. This, in 
turn, causes the pressure plate 1103 to change its axial position relative 
to the leaf springs 1122 which abut the flywheel 1106 in the axial 
direction of the friction clutch. Since the ring 1126 is biased by the 
diaphragm spring 1104, this ring shares the axial movement of the pressure 
plate 1103 toward the flywheel 1106 to an extent which is determined by 
wear (primarily) upon the friction linings 1107. This, in turn, causes the 
projections 1132 of the ring 1126 to move axially and away from the 
corresponding projections 1133 of the leaf springs 1122 through a distance 
which also corresponds to or at least approximates the reduction in the 
thickness of friction linings 1107 due to wear as a result of repeated 
frictional engagement with and disengagement from the friction surfaces of 
the pressure plate 1103 and flywheel 1106. The axial position of the ring 
1126 relative to the pressure plate 1103 remains unchanged during 
engagement of the friction clutch 1101 because the ring is acted upon by 
the diaphragm spring 1104 in a direction toward the pressure plate and the 
displacing device 1134 is self-locking in the course of the clutch 
engaging operation, i.e., the device 1134 acts as a means for "locking" 
the ring 1126 to the pressure plate 1103 during engagement of the friction 
clutch. When the clutch is being disengaged, i.e., when the tips 1104c of 
the prongs 1104b move along their path in the opposite direction, the 
pressure plate 1103 is biased by the leaf springs 1109 which urge the 
pressure plate toward the bottom wall 1102a of the cover 1102. The 
pressure plate 1103 ceases to move axially toward the bottom wall 1102a 
until the leaf springs 1122 engage the cover 1102 and more specifically 
the marginal portion 1123 of thus cover. The extent of movement of the 
prongs 1104b in a direction to disengage the friction clutch 1101 
corresponds to the extent of movement of the pressure plate 1103 away from 
the flywheel 1106, and the axial position of the ring 1126 relative to the 
pressure plate remains unchanged. If the movement of the prongs 1104b in a 
direction to disengage the friction clutch 1101 continues, the pressure 
plate 1103 comes to a halt (i.e., its axial position remains unchanged) 
but the ring 1126 continues to share the axial movement of the adjacent 
portion of the diaphragm spring 1104 (namely of the portion which bears 
upon the end wall 1127 of the ring 1126). The ring 1126 is arrested and no 
longer moves toward the bottom wall 1102a of the cover 1102 when the 
projections 1132 at the end wall 1127 of the ring reengage the projections 
1133 of the leaf springs 1122. Axial shifting of the ring 1126 is effected 
by the wedges 1135 which are biased by the coil springs 1144 to move 
relative to the adjacent wedges 1136 in the circumferential direction of 
the ring 1126 until the projections 1132 again engage and bear against the 
respective projections 1133 on the leaf springs 1122. 
In the friction clutch 1101 of FIGS. 29 to 32, the pressure plate 1103 can 
be disengaged from the adjacent set of friction linings 1107 (i.e., moved 
axially and away from the flywheel 1106) by the stressed leaf springs 
1109. Thus, the leaf springs 1109 always tend to disengage the pressure 
plate 1103 from the clutch disc 1108, i.e., to move the pressure plate 
toward the bottom wall 1102a of the cover 1102. If the diaphragm spring 
1104 continues to perform a movement in a sense to disengage the friction 
clutch, the radially outer portion of the diaphragm spring moves axially 
and away from the ring 1126 because the ring 1126 is arrested and no 
longer moves toward the bottom wall 1102a when its projections 1132 
reengage the projections 1133 of the leaf springs 1122. Such, even very 
slight, disengagement of the diaphragm spring 1104 from the ring 1126 
during disengagement of the friction clutch 1101 is of particular 
advantage for the system including the devices 1117 and 1134. 
The devices 1117 and 1134 ensure that adjustment of the ring 1126 as a 
result of shifting of the wedges 1135 and their ramps 1143 relative to the 
ramps 1142 of the wedges 1136 invariably compensates for wear upon the 
pressure plate 1103, flywheel 1106 and friction linings 1107. This is 
attributable to the fact that the ring 1126 is clamped between the 
adjusting elements (wedges) 1135, 1136 on the one hand, and the leaf 
springs 1122 on the other hand (as seen in the axial direction of the 
friction clutch 1101); this prevents the ring 1126 from performing an 
axial movement greater than that corresponding to wear (primarily) upon 
the friction linings 1107. In addition, the devices 1117 and 1134 ensure 
that, even if the prongs 1104b of the diaphragm spring 1104 cover a 
distance greater than necessary to disengage the friction clutch 1101, or 
if the pressure plate 1103 is caused to perform axial vibratory movements 
relative to the flywheel 1106, the wedges 1135, 1136 do not effect any 
adjustment of the type required to take place in order to compensate for 
wear upon the friction linings 1107. The reason is that the leaf springs 
1122 do not move relative to the pressure plate 1103 and/or vice versa, 
even in the event of a pronounced impact of their stops or projections 
1133 against the marginal portion 1123 of the cover 1102. The reason is 
that the displacing device 1134 is self-locking by way of the projections 
1132. Thus, even if the friction clutch 1101 is disengaged, the leaf 
springs 1122 can be acted upon by forces acting in the axial direction of 
the friction clutch toward the flywheel 1106 and having a magnitude 
exceeding that between the leaf springs 1122 and the pressure plate 1103 
without risking any axial displacement of the pressure plate and leaf 
springs 1122 relative to each other. 
The improved adjusting unit 1116 ensures that, for all practical purposes, 
only a certain part of the characteristic curve of the diaphragm spring 
1104 requires consideration during the entire useful life of the friction 
clutch (i.e., while the wear upon the friction linings 1107 progresses 
from zero to a maximum permissible value). Moreover, the bias of the 
pressure plate 1103 upon the friction linings 1107 in engaged condition of 
the friction clutch is the same irrespective of the extent of wear upon 
the friction linings because the bias of the diaphragm spring 1104 upon 
the pressure plate remains unchanged. This, in turn, renders it possible 
to employ a diaphragm spring 1104 having a degressive characteristic curve 
during actuation of the means for disengaging the friction clutch, 
preferably in combination with a clutch plate or clutch disc 1108 wherein 
the two sets of friction linings 1107 are biased apart by resilient 
segments 1110 or the like. This renders it possible to reduce the 
magnitude of the effective clutch disengaging force to a relatively low 
level and to maintain the disengaging force at such low level during the 
entire useful life of the friction clutch 1101, as long as the 
characteristic curve of the resilient segments 1110 remains at least 
substantially unchanged during the useful life of the friction clutch. 
When the friction clutch is being disengaged, the diaphragm spring 1104 is 
tilted at the seat 1105 whereby the stressing of the resilient segments 
1110 decreases during a certain portion of axial movement of the prongs 
1104b along their path, i.e., during a certain stage of axial movement of 
the pressure plate 1103 away from the flywheel 1106. The resilient 
segments 1110 dissipate energy during the just mentioned stage of movement 
of the pressure plate 1103 away from the flywheel 1106 to thus assist in 
disengagement of the friction clutch. This means that the maximum force 
which is required to disengage the friction clutch 1101 is smaller than 
the theoretical force generated by and attributable to the mode of 
installation of the diaphragm spring 1104 in engaged condition of the 
friction clutch. When the range of resiliency of the segments 1110 is 
exceeded, the friction linings 1107 are released (disengaged from the 
pressure plate 1103 and flywheel 1106) and, due to the degressive 
characteristic curve of the diaphragm spring 1104 during disengagement of 
the friction clutch, the remaining disengaging force which is to be 
applied is much less than that disengaging force which would correspond to 
the installation point or position of FIG. 29. As the disengagement of the 
friction clutch 1101 continues, the magnitude of the disengaging force 
continues to decrease at least until the lowest point of the preferably 
sinusoidal characteristic curve of the diaphragm spring 1104 is reached. 
It is advantageous to design the devices 1117 and 1134 in the friction 
clutch 1101 of FIGS. 28 and 29 in such a way that, when the friction 
clutch is driven, the individual convolutions of the coil springs 1144 in 
the space 1126a of the ring 1126 abut the radially outer sidewall 1131 of 
the ring 1126. Friction between the springs 1144 and the sidewall 1131 
then opposes or completely neutralizes the forces which the springs 1144 
tend to apply in the circumferential direction of the ring 1126, i.e., 
those forces which would tend to shift the wedges 1135 relative to the 
wedges 1136 and to thus compensate for wear upon the friction linings 1107 
at a time when such compensation is not necessary. In other words, the 
springs 1144 act not unlike rigid bodies when the friction clutch 1101 is 
driven by the internal combustion engine of a motor vehicle or by any 
other prime mover, and such behavior of the springs 1144 is attributable 
to friction between their convolutions and the adjacent internal surface 
of the radially outer sidewall 1131 of the ring 1126. In addition, the 
wedges 1135 are also acted upon by centrifugal force which urges them 
against the internal surface of the radially outer sidewall 1131 of the 
ring 1126 so that the wedges 1135 are in frictional engagement with the 
sidewall 1131 and are not likely to move in the circumferential direction 
of the ring 1126 while the friction clutch 1101 rotates and the springs 
1144 are acted upon by centrifugal force. The arrangement may be such that 
the magnitude of centrifugal force acting upon the springs 1144 and wedges 
1135 suffices to prevent any undesirable shifting of these wedges in the 
circumferential direction of the ring 1126 unless the rotational speed of 
the friction clutch 1101 is within the idling RPM range of the internal 
combustion engine provided that the friction clutch is put to use between 
the engine and the variable-speed transmission of a motor vehicle. At such 
time, the springs 1144 are incapable of effecting any shifting of the 
wedges 1135 relative to the adjacent wedges 1136. Thus, the friction 
clutch 1101 can be designed in such a way that any compensation for wear 
upon the friction linings 1107 can take place only when the RPM of the 
engine is within or at least close to the idling RPM. Blocking of 
adjustment to compensate for wear upon the friction linings 1107 during 
certain stages of operation of the motor vehicle which employs the 
improved friction clutch 1101 can also be accomplished only when the 
internal combustion engine or any other prime mover which is used to 
rotate the friction clutch is idle or its RPM is negligible, i.e., when 
the flywheel 1106 does not rotate and does not transmit torque to the 
pressure plate 1103 and cover 1102 or the RPM of the flywheel 1106 is 
minimal. All that is necessary is to carry out corresponding adjustments 
in the design of the displacing device 1134. 
The materials of the wedges 1135, 1136 and of the parts which cooperate 
with these wedges are preferably selected in such a way that the wedges of 
the pairs of wedges 1135, 1136 do not tend to adhere to each other during 
any stage of useful life of the friction clutch 1101, i.e., that adherence 
between the ramps 1142, 1143 of pairs of cooperating wedges 1135, 1136 
cannot rise to a value at which the device 1134 would be incapable of 
compensating for wear upon the friction linings 1107. Undesirable 
adherence of the ramps 1142 to the adjacent ramps 1143 can be prevented by 
coating at least one of the ramps of each pair of wedges 1135, 1136 with a 
suitable friction reducing or preventing (lubricating) material. 
It is further possible to prevent adherence of the ramps 1142 and the 
neighboring ramps 1143 to each other by providing the friction clutch 1101 
with one or more systems or devices which apply to the ramps 1135 an 
axially oriented force in a direction axially of the friction clutch and 
away from the neighboring wedges 1136 in order to break the bonds (if any) 
between the neighboring ramps 1142 and 1143 in response to each 
disengagement of the friction clutch. This ensures that the device 1134 is 
ready to accurately compensate for any and all wear upon the friction 
linings 1107 and, if necessary, also upon the friction surfaces of the 
flywheel 1106 and pressure plate 1103. 
Referring to FIG. 30, the position of the mobile wedge 1135 which is 
illustrated therein relative to the adjacent wedge 1136 departs from the 
initial position, namely from that position which the wedge 1135 assumes 
(as seen in the circumferential direction of the ring 1126) when the wear 
upon the friction linings 1107 is negligible, minimal or nil, for example, 
prior to mounting of the pressure plate 1103 and the cover 1102 on the 
flywheel 1106. At such time, the ring 1126 assumes a position at a minimal 
axial distance from the pressure plate 1103. Expressed otherwise, the 
combined thickness of the pressure plate 1103 and ring 1126 then assumes a 
minimum value. In order to ensure that the wedges 1136 will remain in 
their fully retracted positions (nearest to the end wall 1127 of the ring 
1126) while the cover 1102 and/or the pressure plate 1103 is being 
connected to the flywheel 1106, the wedges 1135 are preferably provided 
with portions (e.g., in the form of recesses or notches 1152 shown in FIG. 
30) which can receive the working ends of suitable retaining or retracting 
tools. Such tools are put to use during assembly of the structure 
including the ring 1126, coil springs 1144 and wedges 1135, 1136 and/or 
during attachment of the cover 1102 and pressure plate 1103 to the 
flywheel 1106 in order to ensure that the mobile wedges 1135 will be fully 
retracted when the friction clutch 1101 is assembled and the wear upon its 
parts 1103, 1106, 1107 is still zero or negligible. It is clear that the 
just discussed tools are removed (disengaged from the wedges 1135 and/or 
1136) when the assembly of the friction clutch 1101 is completed; this 
ensures that the device 1134 is then ready to ensure necessary adjustments 
to compensate for wear upon the parts 1103, 1106 and/or 1107. As can be 
seen in FIGS. 30 and 32, the ring 1126 is provided with elongated slots 
1153 which enable the working ends of one or more retaining tools to enter 
the notches 1152 of the wedges 1135, 1136. For example, the means for 
retracting the wedges 1135 or for maintaining the wedges 1135 in retracted 
positions prior to completed assembly of the friction clutch 1101 can 
comprise one or more turning or rotating tools. The length of the slots 
1153 (which extend in the circumferential direction of the ring 1126) 
should suffice to ensure that the wedges 1135 can be shifted relative to 
the associated wedges 1136 through a distance not less than the maximum 
range of adjustment of wedges 1135 relative to the wedges 1136 for the 
purpose to compensate for maximum wear upon the friction linings 1107. 
When the wedges 1135 are moved to their fully retracted starting positions 
(corresponding to those when the wear upon the linings 1107 is zero), the 
thus retracted wedges 1135 can be maintained in such positions by the leaf 
springs 1122 which secure the ring 1126 in the retracted angular position. 
The self-adjusting connections between the leaf springs 1122 and the 
pressure plate 1103 must be designed in such a way that the shifting or 
displacing force which is required to move the leaf springs 1122 relative 
to the pressure plate 1103 exceeds the resultant of forces acting upon the 
ring 1126 and furnished by the coil springs 1144, i.e., by the springs 
which tend to shift the wedges 1135 relative to the adjacent wedges 1136. 
FIG. 30 shows that the wedges 1135 and their ramps 1143 are separately 
produced parts which are introduced into the space 1126a of and are 
secured to the ring 1126. It is possible to avoid the making of discrete 
wedges 1135 by the simple expedient of properly shaping (deforming, such 
as stamping) the end wall 1127 of the ring 1126, i.e., the wedges 1135 can 
constitute integral parts of (they can be of one piece with) the ring 
1126. The springs 1144 are then designed to turn the ring 1126 (with its 
integral wedges 1135 and/or ramps 1143) relative to the pressure plate 
1103. The other wedges 1136 (or at least the ramps 1142) can be of one 
piece with the pressure plate 1103. Alternatively, the wedges 1136 can be 
produced in a separate step to be thereupon affixed (e.g., welded, glued 
and/or otherwise bonded) to the pressure plate 1103. The thus modified 
friction clutch must employ a ring 1126 with projections corresponding to 
but being much longer than the projections 1132 (as seen in the 
circumferential direction of the ring 1126) in order to ensure that the 
length of the modified projections corresponding to the projections 1132 
will at least match that angular displacement of the ring 1126 which is 
necessary to ensure a full range of automatic adjustments of the axial 
position of the pressure plate 1103 relative to the flywheel 1106 in order 
to compensate for wear upon the parts 1103, 1106 and/or 1107. This ensures 
that an axial limit or stop between the leaf springs 1122 and the ring 
1126 is established and maintained during the entire useful life of the 
thus modified friction clutch. In the just described embodiment of the 
friction clutch (i.e., in that modification of the friction clutch wherein 
the wedges 1135 are of one piece with the ring 1126 and the wedges 1136 
are of one piece with the pressure plate 1103), the angular position of 
the ring 1126 relative to the pressure plate 1103 can be changed from 
without upon completed assembly of the friction clutch. For example, it is 
possible to change the angular position of the ring 1126 in response to 
engagement of its projections 1132 which are made accessible through 
windows or other suitable radially extending openings in the radially 
outer portion of the cover 1102. Such openings or windows can further 
serve to receive the torque transmitting lobes 1125 of the pressure plate 
1103 and/or the leaf springs 1109. 
The adjusting unit 1116 which is shown in FIGS. 28-32 and its 
aforedescribed modifications exhibit the advantage that the novel features 
thereof can be embodied with equal advantage in so-called pull-type 
friction clutches wherein the diaphragm spring has a radially outer 
portion tiltably mounted on the cover or housing and radially inner 
portions bearing upon the pressure plate. A portion of such pull type 
friction clutch 1201 is shown in FIG. 33. A unit 1234 which compensates 
for wear at least upon the friction linings (not shown in FIG. 33) is 
installed between the diaphragm spring 1204 and the pressure plate 1203 
and can be constructed and assembled in a manner as described with 
reference to the embodiment of FIGS. 28-32. The ring 1226 of the 
compensating unit 1234 cooperates with wear detecting or sensing means 
1222 by way of sensor elements 1217. The positions of the wear detecting 
means 1222 relative to the pressure plate 1203 are adjusted in that their 
end portions 1222a engage the housing or cover 1202. The wear detecting 
means 1222 are provided with projections or abutments 1233 which limit the 
extent of axial movability of the pressure plate 1203 during disengagement 
of the friction clutch 1201. In order to ensure satisfactory functioning 
of the unit 1234, the ring 1226 is mounted in such a way that it has 
freedom of at least some axial movability relative to the detecting means 
1222. This can be achieved by establishing a connection 1233a between the 
detecting means 1222a and radially extending portions or arms 1226a of the 
ring 1226, whereby the tips of the arms 1226a have a certain minimal 
freedom of movability relative to the respective detecting means 1222 
and/or vice versa. The arms 1226a can be received in the notches of the 
adjacent detecting means 1222 without any clearance if such arms are 
sufficiently resilient to permit the required axial movements of the ring 
1226 and the detecting means 1222 relative to each other. 
FIG. 34 shows a portion of a friction clutch 1301 wherein the sensor 
elements 1317 extend directly into the main portion of the pressure plate 
1303. The wear detecting means 1322 are provided with stops or heads 1322a 
which cooperate with complementary stops 1323 forming part of the housing 
or cover 1302. The stops 1323 are of one piece with securing means 1302a" 
forming part of a seat 1305 for the diaphragm spring 1304. The illustrated 
securing means 1302a" include prongs or lugs which are of one piece with 
the cover 1302 and extend axially of the friction clutch 1301 through the 
diaphragm spring 1304. The wear compensating device 1334 is disposed 
radially outwardly of the sensor elements 1317 which, in turn, are 
adjacent the circumferentially complete main portion 1304a of the 
diaphragm spring 1304. 
An advantage of the improved friction clutch is that its useful life can be 
prolonged by the simple expedient of employing thicker friction linings, 
i.e., by establishing a longer path for adjustment of the pressure plate 
relative to the counterpressure plate in order to compensate for wear upon 
the friction linings. In addition, the improved friction clutch renders it 
possible to reduce the magnitude of the disengaging force by employing an 
energy storing resilient device (e.g., the diaphragm spring 4 or 1104) 
with a degressive force-to-displacement ratio or characteristic in 
combination with at least one resilient element (such as the segments 10 
or 1110) which opposes the bias of the resilient device that acts upon the 
pressure plate. The at least one resilient device ensures a gradual 
increase or decrease of the torque which can be transmitted by the 
friction clutch and its clutch disc during a portion at least of 
engagement or disengagement of the friction clutch, i.e., during at least 
a portion of movement of the actuating means (such as the prongs 4b and 
their tips 4c) of the clutch engaging and disengaging means along its 
predetermined path. The resilient element is preferably installed in 
series with the diaphragm spring. The design of the improved friction 
clutch is such that the magnitude of the disengaging force can be reduced 
to a surprisingly large extent; moreover, such reduction of the required 
disengaging force exists and remains at least substantially unchanged 
during the entire useful life of the improved friction clutch. In other 
words, if it fluctuates at all, the disengaging force fluctuates within a 
very narrow range. 
A further important advantage of the improved friction clutch is that it 
can employ a diaphragm spring whose distance-to-force ratio is relatively 
steep within the entire operating range. The utilization of such diaphragm 
springs in heretofore known friction clutches would result in highly 
pronounced rise of the disengaging force in response to wear upon the 
friction linings. 
In a friction clutch which is not provided with the improved wear 
compensating or adjusting unit, the point 41 (FIG. 8) on the curve 40 
migrates along the sinusoidal path in a direction toward the maximum 41a. 
As already discussed hereinabove, the point 41 denotes an axial force 
which is generated by the diaphragm spring 4 in the friction clutch 1 of 
FIGS. 1-2 when the friction clutch is engaged. During disengagement of the 
friction clutch, the magnitude of the disengaging force decreases in a 
direction toward and up to the point 41b. In general, the level of 
progress of the disengaging force increases in comparison to the level of 
the progress of disengaging force when the friction linings are devoid of 
wear. Thus, the distance 43 shown in FIG. 8 is shifted in a direction to 
the left toward the position 43a until the point 41 coincides with the 
maximum 41a. The point 44 is then transferred accordingly along the path 
which is denoted by the curve 40. As the wear upon the friction linings 
progresses, the installation point of the curve denoting the magnitude of 
the force of the diaphragm spring in engaged condition of the friction 
clutch migrates from the maximum 41a gradually toward the point 41b, i.e., 
the bias of the diaphragm spring upon the pressure plate in a conventional 
friction clutch decreases at a gradual rate. That force of the diaphragm 
spring which is applied to the pressure plate at the point 41b in the 
diagram of FIG. 8 corresponds to the force which is applied when the wear 
upon the friction linings is nil (note the point 41). As soon as the 
maximum 41a is exceeded, disengagement of the friction clutch first 
entails an increase of the disengaging force, at least during a portion of 
movement of the actuating means (such as the prongs and the tips of prongs 
forming part of the diaphragm spring). When the maximum permissible wear 
upon the friction linings is reached (note the point 41b in the diagram of 
FIG. 8), the magnitude of the disengaging force must increase during each 
and every stage of disengagement of a conventional friction clutch which 
is not equipped with the novel adjusting unit. Such rise of the magnitude 
of disengaging force is observable even if the friction linings of the 
conventional friction clutch cooperate with the resilient segments 10 or 
with a substitute for such resilient segments (as indicated in FIG. 8 by 
the broken line 42a). 
In designing the improved friction clutch, and particularly its adjusting 
unit, it is necessary to take into consideration that, if the friction 
clutch is utilized in a power train receiving torque from the output 
element (such as a crankshaft) of an internal combustion engine in a motor 
vehicle, the output shaft is likely to transmit to the flywheel (such as 
the flywheel) 6 in the friction clutch 1 of FIGS. 1 and 2, at least some 
axial and/or other (such as wobbling) vibratory or stray movements. The 
flywheel transmits such undesirable stray movements to other component 
parts of the friction clutch. This could induce the adjusting unit 16 in 
the friction clutch 1 of FIGS. 1 and 2 (or the adjusting unit of any other 
of the various heretofore described friction clutches) to carry out 
certain adjustments for non-existent wear upon the flywheel, the pressure 
plate and/or the friction linings of the friction clutch. In other words, 
it is necessary to undertake certain steps in order to prevent undesirable 
axial, wobbling and/or other stray movements of the flywheel from 
influencing the adjusting unit. In the friction clutches which are shown 
in FIGS. 1 to 27, i.e., in those which are equipped with a sensor 
corresponding to the sensor 13 in the friction clutch 1 of FIGS. 1 and 2, 
the adjusting force of this sensor must exceed the forces of inertia which 
can influence the sensor. Such inertial forces are the sums of forces due 
to inertia of the main diaphragm spring (the spring 4 in the friction 
clutch 1), of the adjusting member (such as 17) and/or the adjusting 
elements (such as 18 and 24), a certain portion of the mass of the sensor 
(such as 13) and, at least in certain instances, the masses of some 
additional components multiplied by the maximum possible axial 
acceleration of these parts and/or components, all due to axial and/or 
other vibratory or other stray movements of the flywheel in response to 
stray movements of the output element of the prime mover. 
By way of example, and referring to the friction clutch 1001 of FIG. 27 
wherein the sensor 1013 engages the pressure plate 1003, it is also 
necessary to take into consideration the inertia of the pressure plate 
1003. Thus, it is necessary to ensure that the force which is generated by 
the sensor will exceed the sum of forces which act upon the sensor and are 
obtained by multiplying the maximum axial acceleration with the combined 
mass of all parts which act upon the sensor due to their inertia. Such 
inertial forces can exert an undesirable influence, particularly during 
actuation of the friction clutch and also in disengaged condition of the 
friction clutch. 
In the embodiments of the improved friction clutch which are shown in FIGS. 
28 to 34, the wear detecting means and the wear compensating means must 
also be designed with a view to take into consideration those forces which 
develop as a result of inertia of all parts which are set in motion in 
response to axial and/or other vibratory movements which are transmitted 
from the output element of the prime mover to the flywheel of the friction 
clutch. 
All in all, the designer of a friction clutch with built-in wear 
compensating or adjusting means must take into consideration the masses of 
those elements which can be acted upon and can be set in motion in 
response to transmission of axial, rotary, wobbling and/or other stray 
movements from the output element of the prime mover to the flywheel of 
the friction clutch. In the embodiments of FIGS. 28-34, it is particularly 
important to take into consideration the inertia of all such parts which 
influence the operation of the ramps, such as the ramps 1142, 1143 in the 
ring 1126 of the friction clutch 1101 shown in FIGS. 28 and 29. 
FIG. 35 shows a friction clutch which embodies or is mounted on one 
(secondary) mass or flywheel 1403 of a composite (twin) flywheel 1401 
further including a flywheel or primary mass 1402 and a damper 1409 
between the masses 1402, 1403. The primary mass 1402 of the composite 
flywheel 1401 can be connected to the output element (e.g., a crankshaft, 
not shown) of an internal combustion engine and transmits torque to the 
secondary mass or flywheel 1403 corresponding, for example, to the 
flywheel 6 in the friction clutch 1 of FIGS. 1 and 2. The friction clutch 
which embodies or is combined with the secondary flywheel 1403 is denoted 
by the reference character 1404. The friction clutch 1404 further 
comprises a pressure plate 1428 and a torque transmitting clutch disc or 
clutch plate 1405 between the pressure plate 1428 and the secondary mass 
1403. The hub 1405a of the clutch disc 1405 can transmit torque to the 
input element of a variable-speed transmission in the power train of a 
motor vehicle. The shaft of the input element of the transmission is 
indicated at X--X. 
An antifriction bearing 1406 is installed between the masses 1402 and 1403 
of the composite flywheel 1401; this bearing is disposed radially inwardly 
of bolts 1408 or other suitable fasteners which are used to secure the 
primary mass 1402 to the output element of the engine. The primary mass 
1402 has bores or holes 1407 for the shanks 1440a of the fasteners 1408. 
The damper 1409 between the masses 1402, 1403 of the flywheel 1401 
includes energy storing elements in the form of coil springs 1410 acting 
in the circumferential direction of the flywheel 1401 and being confined 
in an annular compartment 1412 constituting the radially outer part of a 
chamber 1411 between the masses 1402 and 1403. The chamber 1411 is at 
least partially filled with a viscous fluid, such as oil, grease or 
another lubricant. 
At least the major part of the primary mass 1402 is constituted by a member 
1413 which is made of a metallic sheet material and includes a 
substantially radially extending flange-like portion 1414 having an axial 
protuberance 1415 which is of one piece therewith and is located radially 
inwardly of the holes or bores 1407 for the fasteners 1408. The 
antifriction bearing 1406 which is shown in FIG. 35 comprises a single row 
of spherical rolling elements 1406a and an inner race 1416 surrounding the 
free end of the axially extending protuberance 1415 of the flange 1414. 
The outer race 1417 of the bearing 1406 is received in a central opening 
provided in the radially innermost portion of the secondary mass 1403; the 
latter resembles a substantially flat disc or washer. 
The radially outermost part of the flange 1414 forming part of the primary 
mass 1402 is of one piece with a first wall 1418 which surrounds at least 
one-half of the compartment 1412 and is welded or otherwise sealingly 
secured to a second wall 1419 surrounding another part of the compartment 
1412. The wall 1418 and/or 1419 can directly or indirectly guide the 
radially outermost portions of convolutions forming part of the energy 
storing elements 1410 in the compartment 1412. The reference character 
1420 denotes a welded seam which connects the walls 1418, 1419 to each 
other radially outwardly of the compartment 1412 and ensures that the 
confined viscous fluid cannot escape from the chamber 1411 under the 
action of centrifugal force when the composite flywheel 1401 receives 
torque from the output element of the engine. 
The compartment 1412 is divided into a series of arcuate sections, one for 
each energy storing element 1410, and such sections are separated by 
partitions which constitute abutments for the adjacent end convolutions of 
the respective energy storing elements 1410. The partitions can be made of 
one piece with the wall 1418 and/or 1419 of the primary mass 1402; they 
may constitute inwardly bent pockets of the walls 1418 and 1419. Such mode 
of making partitions between the energy storing elements 1410 is 
particularly desirable when the parts of the primary mass 1402 are made of 
a ductile metallic sheet material. 
The energy storing elements 1410 are further acted upon by radially 
outwardly extending arms 1421 adjacent the secondary mass 1403. The arms 
1421 also alternate with the energy storing elements 1410, as seen in the 
circumferential direction of the composite flywheel 1401, and cooperate 
with the aforediscussed pockets of the primary mass 1402 to ensure that 
the elements 1410 store energy (or additional energy) whenever the mass 
1402 turns relative to the mass 1403 and/or vice versa. These arms are 
provided on or can constitute integral parts of the housing or cover 1422 
of the friction clutch 1404. As shown, the arms 1421 are of one piece with 
the axially extending portion 1423 of the cover 1422. Each arm 1421 
extends radially outwardly into the compartment 1412 between the ends of 
the two neighboring energy storing elements 1410. The axially extending 
portion 1423 of the cover 1422 has a portion 1423a which extends beyond 
the arms 1421 in a direction toward the mass 1402 and surrounds the mass 
1403. The means for connecting the cover 1422 to the mass 1403 can 
comprise inwardly extending portions (not specifically shown) of the 
portion 1423 and complementary sockets in the periphery of the mass 1403. 
Other connecting means (e.g., in the form of radially extending pins or 
the like) can be used with similar advantage. 
The cover 1422 includes a bottom wall 1426 which extends substantially at 
right angles to the axis X--X constituting the common axis of the clutch 
1404 and the input element of the transmission and is remote from the arms 
1421. This bottom wall is outwardly adjacent a diaphragm spring 1427 which 
acts not unlike a two-armed lever and serves to urge the pressure plate 
1428 axially toward the friction linings 1429 of the clutch disc 1405. The 
projecting portion or portions 1428a of the pressure plate 1428 are 
engaged by the circumferentially complete radially outer main portion of 
the diaphragm spring 1427, and the latter includes radially inwardly 
extending prongs 1427a forming part of actuating means for the friction 
clutch 1404, i.e., of means for engaging and disengaging the clutch. 
FIG. 35 further shows resilient segments 1465 which are disposed between 
the two groups or sets of friction linings 1429 and perform the same 
function as the segments 10 in the friction clutch 1 of FIGS. 1 and 2. 
The chamber 1411 and its compartment 1412 are disposed, at least to a large 
extent, radially outwardly of the secondary mass 1403 of the composite 
flywheel 1401. This renders it possible to position the member 1413 of the 
primary mass 1402 (i.e., of that mass which is to be directly connected 
with the output element of an engine) into immediate or close proximity to 
the secondary mass 1403 in a region radially inwardly of the chamber 1411. 
FIG. 35 shows a relatively narrow clearance 1430 which is established 
between the member 1413 of the primary mass 1402 and the secondary mass 
1403. Such design contributes significantly to compactness of the friction 
clutch 1404, as seen in the direction of the axis X--X, and more 
particularly of the aggregate including the friction clutch 1404 proper, 
the composite flywheel 1401 and the clutch disc 1405. 
The chamber 1411 is sealed by an annular sealing element 1431 which is 
installed between the radially inner portion of the wall 1419 (i.e., of 
the primary mass 1402) and the axially extending portion 1423 of the cover 
1422. 
The aforementioned clearance 1430 between the member 1413 of the primary 
mass 1402 and the secondary mass 1403 can be utilized to ensure desirable 
cooling of the composite flywheel 1401. This is achieved by inducing one 
or more currents of cool atmospheric air to flow through the clearance 
when the aforementioned aggregate or assembly is in actual use, i.e., when 
the output element of the engine drives the primary mass 1402 and the 
latter drives the secondary mass 1403 through the damper 1409 including 
the energy storing elements 1410 in the compartment 1412 of the chamber 
1411. The means for cooling the flywheel 1401 further comprises passages 
or channels 1433 which extend through the secondary mass 1403 radially 
inwardly of a friction surface 1432 which is engageable by the adjacent 
set of friction linings 1429 when the friction clutch 1404 is engaged. The 
channels 1433 communicate with the clearance 1430. The cooling action is 
further enhanced by the provision of additional channels 1435 which extend 
axially through the secondary mass 1403 and are disposed radially 
outwardly of the friction surface 1432. The channels 1435 communicate with 
the clearance 1430, the same as the channels 1433. The channels 1433 
supply cool atmospheric air into the radially inner portion of the 
clearance 1430, and such air then flows radially outwardly to cool the 
composite flywheel 1401 and to leave the clearance 1430 through the 
channels 1435. These channels can admit the atmospheric air into the cover 
1422 which is provided with outlets to permit escape of heated air into 
the surrounding atmosphere. 
The secondary mass 1403 is provided with holes or bores 1434 which are 
disposed radially inwardly of the channels 1433 and are aligned with the 
holes or bores 1407 to permit introduction of the fasteners 1408 which 
serve to affix the primary mass 1402 to the output element (e.g., a 
crankshaft) of an engine. In addition, the holes or bores 1434 can also 
promote circulation of air in the clearance 1430, i.e., they can 
contribute to more satisfactory cooling of the composite flywheel 1401. 
A further sealing element 1436 is disposed in the clearance 1430 to seal 
the latter from the radially innermost portion of the annular chamber 1411 
for the supply of viscous fluid and for the energy storing elements 1410 
of the damper 1409. The sealing element 1436 can include or constitute a 
membrane or a diaphragm spring. 
The wall 1418 of the primary mass 1402 is provided with a starter gear 1439 
which is preferably welded thereto. 
The composite flywheel 1401 including the masses 1402, 1403 and the group 
including the friction clutch 1404 and the clutch disc 1405 together 
constitute a preassembled module A which is or can be assembled at the 
manufacturing plant and can be put to storage or shipped to a maker of 
motor vehicles to be affixed to the output element of an engine by the 
fasteners 1408 or in any other suitable way. The assembly of the module A 
at the plant contributes significantly to lower cost of the improved 
aggregate, to lower cost of its storage and shipment, and to lower cost of 
its attachment to the output element of an engine. In order to assemble 
the module A, the friction clutch 1404 is assembled with the secondary 
mass 1403 and with the clutch disc 1405 in a first step. The thus obtained 
subassembly including the components 1403, 1404 and 1405 is thereupon 
assembled with the primary mass 1402 by placing the member 1413 of the 
primary mass next to the secondary mass 1403 so that the masses 1402, 1403 
are coaxial with one another. This takes place before the wall 1419 is 
affixed (welded) to the wall 1418 of the primary mass 1402. The wall 1419 
surrounds the axially extending portion 1423 of the cover 1422 and is 
welded (at 1420) to the wall 1418 in a next following step. Of course, the 
energy storing elements 1410 are inserted into the compartment 1412 of the 
chamber 1411 prior to welding of the walls 1418, 1419 to each other. 
The antifriction bearing 1406 is installed between the masses 1402, 1403 in 
automatic response to proper positioning of the member 1413 of the mass 
1402 relative to the mass 1403; such bearing is installed first on the 
axially extending protuberance 1415 of flange 1414 of the member 1413. The 
fasteners 1408 are inserted into the holes 1407 of the portion 1414a of 
the flange 1414 before the masses 1402, 1403 are angularly movably coupled 
to each other by the damper 1409. Each fastener 1408 can constitute a 
hexagon socket screw, i.e., a screw with a polygonal socket 1440 in its 
head. The initial positions of the fasteners correspond to that of the 
fastener 1408 shown in the lower half of FIG. 35. It is preferred to 
provide means for yieldably holding the shanks 1440a of the fasteners 1408 
in the axial positions corresponding to that of the shank forming part of 
the fastener 1408 shown in the lower half of FIG. 35. The holding means 
prevent accidental displacement or loss of the fasteners 1408 and ensure 
that the shanks 1440a of these fasteners are maintained in optimum 
positions for introduction into complementary tapped bores or holes of the 
output element of the engine. 
The clutch disc 1405 is centered between the pressure plate 1428 of the 
friction clutch 1404 and the friction surface 1432 of the secondary mass 
1403 of the composite flywheel 1401 and is maintained in such position 
while the module A is in storage or in transport to the automobile 
assembly plant. The angular position of the clutch disc 1405 in the module 
A is such that its holes or bores 1443 are aligned with the holes or bores 
1434 in the secondary mass 1403; this renders it possible to introduce the 
working end of a tool (e.g., a device analogous to a screwdriver) into the 
sockets 1440 in the heads of fasteners 1408 in order to drive the shanks 
1440a of such fasteners into the complementary tapped bores or holes in 
the output element of the engine. The tool can further extend through 
aligned holes or bores 1444 which are provided in the prongs 1427a of the 
diaphragm spring 1427 and communicate with the slots between neighboring 
prongs. The diameters of the holes or bores 1443 are smaller than the 
diameters of the heads of the fasteners 1408 so that, once installed in a 
manner as shown in the lower part of FIG. 35, the fasteners 1408 of a 
module A cannot become lost or misplaced because they are confined in 
optimum positions for attachment to the output element of an engine in a 
motor vehicle. The openings 1444 in the prongs 1427a of the diaphragm 
spring 1427 can constitute simple recesses or notches; such recesses or 
notches communicate with the slots between the respective prongs 1427a to 
provide room for introduction of the aforediscussed tool which must also 
pass through the holes 1443 and into the holes 1434 in order to enter the 
sockets 1440 in the heads of the respective fasteners 1408. 
It is often preferred to distribute the tapped holes or bores in the output 
element of the engine and the holes or bores 1407 in the member 1414 of 
the primary mass 1402 in such a way that the mass 1402 can be affixed to 
the output element in a single angular position, i.e., the holes 1407 need 
not be equidistant from each other. The dimensions of the openings 1434, 
1443 and 1444 are selected in such a way that they permit the working end 
of a tool to engage the heads of the fasteners 1408, one after the other, 
even if the holes 1434 are uniformly distributed in the secondary mass 
1403, the holes 1443 are uniformly distributed in the clutch disc 1405, 
and the holes 1444 are uniformly distributed in the pronged portion of the 
diaphragm spring 1427. The working end of the tool has a shape such that 
it can be non-rotatably received in the preferably hexagonal socket 1440 
in the head of a fastener 1408. 
The assembly of a module A at the manufacturing plant contributes 
significantly to convenience, simplicity and lower cost of installation of 
the aggregate (including the composite flywheel 1401, the friction clutch 
1404 and the clutch disc 1405) in a motor vehicle. This will be readily 
appreciated since the making of the module A renders it possible to 
dispense with a number of time-consuming operations which are necessary to 
install heretofore known friction clutches in automotive vehicles. For 
example, the clutch disc 1405 is properly centered in the module A so that 
no centering of the clutch disc is needed immediately prior to or during 
attachment of the composite flywheel 1401 to the output element of the 
engine. Furthermore, the clutch disc 1405 is already installed between the 
secondary mass 1403 and the pressure plate 1428 at the time the secondary 
mass 1403 is to be coupled to the primary mass 1402 by the bearing 1406 
and the damper 1409, and the friction clutch 1404 is properly attached to 
the output element as soon as the latter is connected with the primary 
mass 1402 by fasteners 1408. Still further, it is no longer necessary to 
employ a centering mandrel, to center the clutch disc 1405 relative to the 
pressure plate 1428 at the motor vehicle assembly plant, to select and 
insert the fasteners 1408, to connect the friction clutch 1404 with the 
composite flywheel 1401 and/or to extract a centering mandrel during or 
subsequent to attachment of the friction clutch to the engine. 
The friction clutch 1404 is provided with an adjusting unit 1445 which is 
or can be identical with or analogous to any one of the adjusting units 
shown in and described with reference to FIGS. 1 through 27. The adjusting 
unit 1445 includes a sensor 1446 (e.g., in the form of a diaphragm spring 
corresponding, for example, to the spring 13) and an annular member 1447 
corresponding, for example, to the member 17 in the friction clutch 1 of 
FIGS. 1 and 2. 
It is normally preferred, primarily for the purpose of reducing the cost, 
to establish a permanent connection between the cover 1422 and the 
secondary mass 1403. Such permanent connection can be established by 
bonding (such as welding) or by deformation of selected portions of the 
mass 1403 and/or cover 1422 so that the separation of these parts would 
involve at least partial destruction (such as extensive deformation) of 
the cover and/or of the secondary mass. The establishment of such 
permanent connection renders it possible to avoid the use of screws, bolts 
and/or other threaded or other fasteners. Since the aggregate including 
the twin-mass flywheel 1401, the clutch disc 1405 and the friction clutch 
1404 is designed to remain fully assembled during its entire useful life, 
i.e., until the wear upon the friction linings 1429 becomes excessive, 
there is no urgent need to establish a readily separable connection 
between these parts or to establish a connection which would permit 
repeated assembly and dismantling of the aggregate. In spite of the 
absence of means for permitting repeated dismantling and assembly of the 
aggregate which is shown in FIG. 35, such aggregate functions 
satisfactorily during its entire useful life because the adjusting unit 
1445 compensates for wear upon the friction linings 1429 but preferably 
also for wear upon one or more additional parts such as the secondary mass 
1403 and/or the pressure plate 1428. The dimensions of the freshly 
installed friction linings 1429 can be selected with a view to ensure that 
they do not become useless due to excessive wear prior to expiration of 
the anticipated useful life of the aggregate. As a rule, the useful life 
of the aggregate will be selected to at least match the anticipated life 
span of the motor vehicle in which the aggregate is being put to use. 
Twin-mass flywheels which can be used in the improved aggregate, e.g., in a 
manner as shown in FIG. 35, are disclosed, for example, in published 
German patent applications Serial Nos. 37 21 712, 37 21 711, 41 17 571, 41 
17 582 and 41 17 579. The features which are disclosed in the just 
enumerated published patent applications can be combined with the features 
of the improved friction clutch and/or with the features of the improved 
aggregate in a number of different ways. By way of example only, the 
aforementioned published German patent application Serial No. 41 17 579 
discloses several manners of establishing a connection between the housing 
or cover and a flywheel in such a way that the connection cannot be 
terminated without at least partial destruction of the flywheel and/or 
housing. 
The utilization of an adjusting device 1445 in an aggregate which employs a 
composite flywheel for transmission of torque from a prime mover to the 
cover and/or pressure plate of a friction clutch is advisable and 
advantageous on the additional ground that the damper 1409 between the 
masses 1402, 1403 can prevent the transmission to the mass 1403 (i.e., to 
the counterpressure plate of the friction clutch 1404) of a number of 
stray movements which would be likely to adversely influence the operation 
of the adjusting unit 1445. The damper 1409 is preferably installed 
radially outwardly of the friction linings 1429 and radially outwardly of 
the friction surface 1432 on the secondary mass 1403 and/or pressure plate 
1403. In a composite flywheel of the type shown in FIG. 35, the friction 
diameter of the clutch disc 1405 should be smaller than in conventional 
friction clutches which renders it necessary to increase the biasing force 
in dependency on the ratio of average friction radii in order to be in a 
position to transmit a predetermined engine torque. If a conventional 
friction clutch (without the adjusting unit 1445) were used, this would 
necessitate an increase of the disengaging force. By employing in the 
aggregate of FIG. 35 a friction clutch with an adjusting unit 1445 (e.g., 
an adjusting unit of the type described with reference to FIGS. 1 to 7a), 
it is now possible to reduce the disengaging force and to thus avoid an 
increase of disengaging force above that which is required in a 
conventional friction clutch. In fact, it is now possible to reduce the 
disengaging force below that which must be applied in a conventional 
friction clutch in spite of the fact that the adjusting unit 1445 renders 
it possible to compensate for wear during the entire useful life of the 
friction clutch and/or of the structure (such as a motor vehicle) in which 
the improved friction clutch is put to use. 
Referring to FIGS. 36 and 37, there is shown a torque transmitting 
arrangement or assembly 1501 comprising a counterpressure plate 1503 which 
is non-rotatably connectable to the output element K (e.g., a crankshaft) 
of an internal combustion engine, and a friction clutch 1504 connected to 
the plate 1503 in such a way that a clutch plate or clutch disc 1505 is 
disposed between the plate 1503 and a pressure plate 1528 of the friction 
clutch 1504. The hub of the clutch disc 1505 transmits torque to the input 
element (e.g., an externally splined shaft) of a variable-speed 
transmission in the power train between the friction clutch 1504 and the 
wheels of a motor vehicle. The axis of the input element of the 
transmission is shown at X--X. 
The friction clutch 1504 comprises a housing or cover 1522 having an 
axially extending marginal portion 1523 which surrounds the pressure plate 
1528 and the friction linings 1529 of the clutch disc 1505. The free end 
1523a of the marginal portion 1523 (the latter can be said to resemble a 
relatively short sleeve or tube) surrounds the counterpressure plate 1503 
and is non-rotatably connected thereto. For example, the free end 1523a 
can be provided with radially inwardly extending protuberances, lugs or 
like parts 1524 which extend into complementary sockets or recesses of the 
counterpressure plate 1503 to ensure that this plate and the cover 1522 
rotate as a unit. However, it is also possible to connect the cover 1522 
with the counterpressure plate 1503 in any one of a number of other ways; 
for example, these parts can be welded to each other or the connections 
between these parts can include threaded fasteners, pins, studs, posts or 
like parts preferably extending in the radial direction of the 
counterpressure plate 1503 and of the marginal portion 1523 of the cover 
1522. The just discussed connecting means preferably also serve to 
accurately center the counterpressure plate 1503 and the cover 1522 
relative to each other. 
The cover 1522 comprises an annular section or bottom wall 1526 which 
extends radially inwardly of the marginal portion 1523 and is outwardly 
adjacent a diaphragm spring 1527 which acts not unlike a two-armed lever 
and serves to bias the pressure plate 1528 toward the adjacent set of 
friction linings 1529 forming part of the clutch disc 1505. The radially 
outermost part of the circumferentially complete main portion of the 
diaphragm spring 1527 can bear against the projecting portion or portions 
of the pressure plate 1528, and a radially inner part of such main portion 
is tiltably mounted at the inner side of the bottom wall 1526 by a seat. 
The radially inwardly extending prongs 1527a of the diaphragm spring 1527 
constitute the actuating means of the means for engaging and disengaging 
the friction clutch 1504. When the clutch 1504 is engaged, the radially 
outermost part of the main portion of diaphragm spring 1527 causes the 
pressure plate 1528 to bear against the adjacent set of friction linings 
1529 and also causes the other set of friction linings 1529 to bear 
against the friction surface of the counterpressure plate 1503. The means 
for engaging and disengaging the friction clutch 1504 further comprises a 
conventional bearing or a pedal (similar or analogous to a gas pedal in a 
motor vehicle) which must be actuated by the driver in order to move the 
prongs 1527a along their predetermined path and to thus effect the 
engagement or disengagement of the friction clutch 1504. 
The means for transmitting torque between the pressure pate 1528 and the 
cover 1522 of the friction clutch 1504 which is shown in FIG. 36 comprises 
leaf springs 1521 each having a first end portion affixed to the cover 
1552 and a second end portion affixed to the pressure plate 1528. It is 
presently preferred to employ rivets 1521a or analogous fasteners as a 
means for connecting the leaf springs 1521 to the pressure plate 1528 
and/or to the cover 1522. As can be seen in the upper part of FIG. 36, the 
rivets 1521a are preferably of the type known as blind rivets; in FIG. 35, 
one such blind rivet is denoted by the character 1490. 
The friction clutch 1504, i.e., the torque transmitting arrangement or 
assembly 1501, comprises an adjusting unit 1545 which is analogous to the 
adjusting units of friction clutches shown in FIGS. 1 to 27 and includes a 
diaphragm spring or sensor 1546 and an annular adjusting member 1547. The 
adjusting unit 1545 serves to compensate for wear upon the pressure plate 
1528 and upon the counterpressure plate 1503 but particularly or primarily 
for wear upon the friction linings 1529. 
The adjusting unit 1545 includes ramps which are provided directly in the 
annular member 1547 and are designed in such a way that they establish air 
transmitting passages 1547a. The member 1547 is located at the inner side 
of the bottom wall 1526 of the cover 1522, and the passages 1547a extend 
in the direction of rotation of the friction clutch 1504. Such passages 
promote desirable cooling of the friction clutch 1504 when the 
counterpressure plate 1503 is rotated by the output element K of the 
engine because the passages induce the flow of currents of cool air. This 
reduces the thermal stresses upon the annular member 1547 which can be 
made of a suitable plastic material. The annular member 1447 of the 
adjusting unit 1445 of the friction clutch 1404 shown in FIG. 35 can be 
constructed and configurated in the same way as the annular member 1547. 
The means for affixing the counterpressure plate 1503 to the output element 
K of the engine comprises an axially elastic coupling element 1550 which 
enables the plate 1503 to perform limited axial movements relative to the 
output element and/or vice versa. The illustrated coupling element 1550 is 
a disc having a stiffness or rigidity such that it can effectively damp 
axial, wobbling, angular and/or other stray movements which the output 
element K would transmit to the friction clutch 1504 and which could 
interfere with accuracy of adjustments carried out by the unit 1545. The 
coupling element 1550 need not damp any and all stray movements; however, 
its damping action should be sufficient to ensure that the unit 1545 can 
properly adjust the position of the pressure plate 1528 in dependency upon 
the extent of wear on certain parts of the friction clutch 1504 and the 
aggregate 1501, especially in dependency on the wear upon the friction 
linings 1529. Furthermore, the elastic coupling element 1550 ensures 
proper operation of the friction clutch 1504 by ensuring proper operation 
of the adjusting unit 1545. Otherwise stated, the coupling element 1550 
should constitute a barrier which is capable of transmitting torque from 
the output element K of the engine to the counterpressure plate 1503 but 
is also capable of shielding the counterpressure plate 1503 and the 
friction clutch 1504 from any such axial, angular and/or other stray 
movements of the output element K which could adversely affect the 
operation of the friction clutch 1504 and particularly the operation of 
the adjusting unit 1545. In the absence of the coupling element 1550, or 
of a functional equivalent of this coupling element, the unit 1545 would 
be likely to carry out unnecessary axial adjustments of the position of 
the pressure plate 1528 relative to the counterpressure plate 1503 or not 
to carry out such adjustments when they are warranted in view of the 
extent of wear upon the friction linings 1529. Unnecessary adjustments by 
the unit 1545 would be attributable primarily to the mass of various parts 
of the aggregate 1501 and to acceleration of such mass due to vibration of 
the output element K and (in the absence of the elastic coupling element 
1550) of various parts of the friction clutch 1504. Alternatively, the 
relatively simple adjusting unit 1545 would have to be replaced with a 
much more complex adjusting unit, namely a unit designed with a view to 
take into consideration a host of additional variables including the 
inertia-induced forces acting upon the component parts of the adjusting 
unit. Moreover, all such inertia-induced forces would have to be properly 
related to each other in order to ensure that the thus modified adjusting 
unit would respond only and alone to signals pertaining to the extent of 
wear upon the pressure plate 1528, the counterpressure plate 1503 and/or 
the friction linings 1529. As a rule, a thus modified adjusting unit (to 
be used in lieu of the unit 1545 in the absence of the coupling element 
1550) would require a number of additional parts and its space 
requirements would greatly exceed those of the unit 1545. 
The adjusting unit 1545 of FIG. 36 operates between the cover 1522 and the 
pressure plate 1528 of the friction clutch 1504. However, it is equally 
possible to equip the aggregate 1501 with a friction clutch of the type 
shown in FIGS. 28 to 34, i.e., with a friction clutch wherein the 
adjusting means serving to compensate for wear upon the friction linings 
is disposed between the diaphragm spring and the pressure plate which is 
biased by the diaphragm spring. 
The radially outer portion of the counterpressure plate 1503 in the 
aggregate 1501 of FIG. 36 is fixedly connected to the elastic coupling 
element 1550 by bolts 1551 or analogous threaded fasteners. For example, 
the bolts 1551 can be replaced with blind rivets of the type shown in FIG. 
35, as at 1490, to connect leaf springs with the pressure plate 1428 of 
the friction clutch 1404. A narrow radially extending gap 1552 is 
established between the neighboring surfaces of the counterpressure plate 
1503 and the coupling element 1550 radially inwardly of the fasteners 
1551; the width of this gap (as measured in the direction of the axis 
X--X) determines the maximum amplitude of axial stray movements which can 
be damped by the element 1550 when the aggregate 1501 of FIG. 36 is in 
use. More specifically, the width of the gap 1552 determines the maximum 
amplitude of those axial movements which are directed from the output 
element K toward the counterpressure plate 1503. The width of the gap 1552 
further determines the extent of maximum movability of the friction clutch 
1504 and counterpressure plate 1503 toward the output element K. As a 
rule, the central portion of the counterpressure plate 1503 does not 
contact the coupling element 1550 if the engine functions properly. 
The counterpressure plate 1503 is a ring which surrounds an axial 
protuberance 1553 of a washer-like member 1554; the latter is fixedly 
secured to the central portion of the elastic coupling element 1550 and 
can serve as a means for centering the element 1550 on a coaxial stub-like 
tubular projection 1555 of the output element K. The radially inner 
portion of the element 1550 is clamped between a front end face 1557 of 
the output element K and the centering member 1554. 
The axial protuberance 1553 of the centering member 1554 has radially 
outwardly extending portions 1558 which constitute stops in that they 
limit the extent of movability of the counterpressure plate 1503 axially 
and away from the central portion of the elastic coupling element 1550 and 
output element K. To this end, the projecting portions or stops 1558 
extend behind the central portion of the plate 1503, i.e., such central 
portion of the plate 1503 is located between the central portion of the 
element 1550 and the stops 1558. A narrow slot or clearance 1559 is 
normally established between the stops 1558 and the central portion of the 
plate 1503, and the width of this clearance 1559 can equal or approximate 
the width of the gap 1552. 
The surface surrounding the central opening of the counterpressure plate 
1503 can be slipped onto the centering member 1554 without any or with a 
minimum of play, i.e., the plate 1503 can be mounted on the member 1554 
without any or with a minimum of radial play but is movable axially 
thereon to the extent which is determined by the gap 1552 and the 
clearance 1559. In other words, the centering member 1554 can be said to 
constitute a guide which confines the counterpressure plate 1503 to 
movements in the direction of the axis X--X. However, it is equally within 
the purview of the invention, and often preferable, to mount the radially 
inner portion of the ring-shaped counterpressure plate 1503 on the portion 
1553 of the centering member 1554 with at least some radial play to thus 
ensure that, in normal operation of the aggregate 1501 (and assuming that 
the operation of the engine including the output element K is 
satisfactory), the counterpressure plate 1503 need not be in any contact 
with the centering member 1554 and/or its portion 1553 and/or the 
projections 1558 and/or the central portion of the elastic coupling 
element 1550. 
It is further within the purview of the invention to provide the aggregate 
1501 with additional means for preventing the transfer of stray movements 
between the output element K and the counterpressure plate 1503 or to use 
such additional means in lieu of the element 1550. For example, the 
additional preventing means can be designed to damp any such stray 
movements which cannot be damped and/or otherwise counteracted by the 
coupling element 1550 to thus even further ensure reliable operation of 
the adjusting unit 1545. Such additional preventing means can be designed 
to destroy energy which is attributable to vibratory and/or other stray 
movements of the output element K, e.g., in a manner as shown in FIG. 37, 
namely by relying on friction. 
FIG. 37 shows that the radially innermost portion of the counterpressure 
plate 1503 and the external surface of the annular portion 1553 of the 
centering member 1554 are separated from each other by a further damper 
1560. For example, the damper 1560 can consist of or can utilize a ring 
which is undulated in the circumferential direction so that its 
undulations extend radially. The ring of the damper 1560 can be installed 
in radially stressed condition to establish friction between its external 
surface and the surface surrounding the central opening of the 
counterpressure plate 1503 whenever the output element K causes the member 
1554 and its portion 1553 to perform stray movements in the direction of 
the axis X--X. In other words, the ring 1560 can prevent the transfer of 
stray movements from the output element K to the counterpressure plate 
1503 or reduces the amplitude of such movements to an acceptable minimum. 
It is possible to utilize a friction generating ring 1560 in the form of a 
split ring. 
The radially outermost portion of the elastic coupling element 1550 carries 
a starter gear 1561 which can be welded or otherwise affixed thereto. 
The coupling element 1550, the counterpressure plate 1503, the clutch disc 
1505 and the friction clutch 1504 can be assembled into a module 
(corresponding to the module A shown in FIG. 35) which can be assembled at 
the manufacturing plant for convenient storage, shipment to an automobile 
assembling plant, and mounted on the output element K of an engine with 
substantial savings in space, initial cost and assembly cost. The 
fasteners 1556 which are shown in FIG. 36 and serve to secure the 
centering member 1554 and the coupling element 1550 to the output element 
K can constitute hexagon socket screws or bolts. As already described with 
reference to FIG. 35, such fasteners can be installed in the 
aforediscussed module in such a way that they cannot be lost and are 
maintained in optimum positions for attachment to the output element K of 
the engine. 
The clutch disc 1505 of the aggregate 1501 which is shown in FIG. 36 is 
installed between and is centered relative to the pressure plate 1528 of 
the friction clutch 1504 and the counterpressure plate 1503 of the 
aggregate 1501. Moreover, the openings or holes 1562 which are provided in 
the clutch disc 1505 are in at least partial alignment with openings 1564 
in the pronged radially inner portion 1527a of the diaphragm spring 1527 
in order to permit the penetration of the working end of a tool 1563 into 
the polygonal sockets in the heads of the fasteners 1556 when it becomes 
necessary to drive the shanks of such fasteners into complementary tapped 
bores or holes in the output element K. The illustrated clutch disc 1505 
comprises an input portion including the friction linings 1529, an output 
portion including the aforementioned hub which can be non-rotatably 
slipped onto the input element of a transmission, and a suitable damper 
employing coil springs or otherwise configurated energy storing elements 
disposed between the input and output portions; the holes 1562 are 
disposed radially inwardly of the damper between the input and output 
portions of the clutch disc 1505 which is shown in FIG. 36. The holes 1564 
in the pronged portion 1527a of the diaphragm spring 1527 are optional, 
i.e., such holes or bores are necessary only if the tool 1563 cannot pass 
through the slots between the neighboring prongs of the diaphragm spring 
1527. The extent of alignment between the holes or bores 1564, the holes 
or bores 1562 and the heads of the fasteners 1556 should suffice to ensure 
that the working end of the tool 1563 will be capable of entering the 
sockets in the heads of the fasteners 1556 even if the holes which are 
provided in the central portion of the elastic coupling element 1550 to 
permit the shanks of the fasteners to pass therethrough are not exactly 
equidistant from each other. As already described with reference to FIG. 
35, such unequal distribution of holes in the coupling element 1550 and in 
the output element K is often desirable in order to ensure that the 
counterpressure plate 1503 can be mounted on the output element K in a 
single predetermined angular position of these parts relative to each 
other. 
As already described with reference to the previously discussed embodiments 
of the present invention, the adjusting unit 1545 enables the friction 
clutch 1504 to operate satisfactorily during its entire useful life. This 
is due to the fact that the unit 1545 can compensate at least for wear 
upon the friction linings 1529 of the clutch disc 1505. Moreover, the 
adjusting unit 1545 renders it possible to permit the utilization of a 
diaphragm spring 1527 which is best suited to ensure that the magnitude of 
the force acting upon the pressure plate 1528 to clamp the friction 
linings 1529 between the friction surfaces of the plates 1503, 1528 
remains within an optimal range for a long interval of time, particularly 
until the wear upon the linings 1529 has progressed to an extent which 
warrants discarding of the aggregate 1501. The diaphragm spring 1527 is 
preferably designed and mounted in such a way that it must merely furnish 
a force which is necessary to ensure adequate biasing of the pressure 
plate 1528 for the purpose of transmitting the desired torque from the 
clutch disc 1505 to the input element of the variable-speed transmission 
in the power train of a motor vehicle. The adjusting unit 1545 ensures 
proper positioning of the diaphragm spring 1527 during the entire life 
span of the aggregate 1501, i.e., it ensures that the bias of the 
diaphragm spring 1527 upon the pressure plate 1528 is satisfactory and 
practically unchanged whenever the friction clutch 1504 is engaged during 
the entire life span of the friction clutch. 
The clutch disc 1505 further comprises resilient segments 1565 which 
constitute a means for gradually reducing the torque which is transmitted 
by the clutch disc 1505 during a portion of movement of the prongs 1527a 
along their path to disengage the friction clutch 1504. Furthermore, the 
segments 1565 ensure a gradual increase of torque which can be transmitted 
from the clutch disc 1505 to the variable-speed transmission during 
engagement of the friction clutch 1504, i.e., while the prongs 1527a of 
the diaphragm spring 1527 are caused to move in the opposite direction. 
This, in turn, renders it possible to reduce the magnitude of the force 
which is necessary to disengage the friction clutch 1504 and to ensure a 
more satisfactory variation of such force in the course of the actual 
disengaging operation. Thus, a desired variation of clutch disengaging 
forces can be achieved by the simple expedient of properly relating the 
forces which are generated by the resilient segments 1565 (or equivalents 
of such segments) and the diaphragm spring 1527, i.e., by properly 
relating the force-to-displacement ratios of such resilient means. This 
renders it possible to optimally design the elastic coupling element 1550, 
i.e., to ensure that the element 1550 will damp any and all stray 
movements which would be likely to adversely influence the operation of 
the adjusting unit 1545. As mentioned above, such stray movements can 
include axial wobbling, bending, angular, tilting and/or other movements 
which are carried out by the output element K and should not be 
transmitted to the counterpressure plate 1503. The magnitude of 
disengaging forces acting upon the coupling element 1550 is minimal. Thus, 
the forces which are required to disengage the friction clutch 1504 can be 
taken up by the element 1550 without any appreciable axial displacement of 
the aggregate 1501. 
The elastic coupling element 1550 can be designed and mounted to shield the 
counterpressure plate 1503 (and hence the adjusting unit 1545) from a 
number of stresses which could result in unintentional or unnecessary 
adjustment of the distance of the pressure plate 1528 from the 
counterpressure plate 1503. It is particularly important to ensure that 
the coupling element 1550 is capable of counteracting the transmission of 
axial and wobbling movements of the output element K to the friction 
clutch 1504. As concerns the construction and mounting of the elastic 
coupling element 1550, reference may also be had to published European 
patent applications Serial Nos. 0 385 752 and 0 464 997 as well as to SAE 
Technical Paper No. 9 003 91. Disclosures of the two European patent 
applications and of the Technical Paper are incorporated herein by 
reference. 
The coupling element 1550 is particularly effective in preventing 
undesirable adjustments attributable to vibratory movements of the 
counterpressure plate 1503 and/or diaphragm spring 1527. Any unintentional 
adjustments of the diaphragm spring 1527, i.e., any adjustments which are 
not necessary to compensate for wear upon the friction linings 1529 but 
are attributable to axial, wobbling and/or other stray movements of the 
output element K, could result in an undesirable reduction of the bias of 
the diaphragm spring upon the pressure plate 1528 below an acceptable 
minimum and would prevent the friction clutch 1504 (and its clutch disc 
1505) from transmitting torques of desired magnitude. 
The aforediscussed design of the improved friction clutch renders it 
possible to maintain the disengaging force at a low value in spite of a 
reduction of the outer diameter of the friction linings and the resulting 
need to increase the bias of the diaphragm spring or its equivalent(s) 
upon the pressure plate. Since the disengaging force is reduced, the 
stressing of the bearing (such as the bearing 1406 in FIG. 35) is less 
pronounced. Thus, it is possible to employ a less expensive antifriction 
bearing and/or a bearing whose space requirements are low. 
Still another advantage of the improved friction clutch and/or of an 
aggregate which employs such friction clutch and/or of a driving unit 
which employs the improved friction clutch and/or the improved aggregate 
is that compensation for wear entails a pronounced lengthening of the 
useful life of the friction clutch. This renders it possible to avoid 
frequent (or any) replacement of parts which are subject to wear, 
particularly the clutch disc 1405. This, in turn, brings about the 
aforediscussed advantage that it is now possible to establish a permanent 
connection between the counterpressure plate and the cover of the friction 
clutch, i.e., a connection whose termination necessitates at least partial 
destruction of at least one of the interconnected parts. Such connection 
can include that which is shown in FIG. 36 and/or a connection which 
employs rivets, welded seams or the like. The establishment of a permanent 
or practically permanent connection is particularly desirable and 
advantageous when the dimensions of the space which is available for the 
improved friction clutch and/or the improved aggregate and/or the improved 
driving unit are small or extremely small, e.g., in a compact motor 
vehicle. Thus, even relatively small reductions of space requirements 
(such as avoiding the use of screws or bolts whose heads would project 
radially outwardly beyond the cover 1422 and/or beyond the composite 
flywheel 1401) are important to ensure that the friction clutch can be 
used in a particular series of motor vehicles. The construction which is 
shown in FIG. 35, as well as the construction which is shown in FIG. 36, 
ensures that, with the exception of the starter gear 1561, the radially 
outermost part of the composite flywheel 1401 or the radially outermost 
part of the cover 1522 determines the maximum space requirements of the 
improved aggregate or driving unit because the means for connecting the 
cover to the counterpressure plate 1403 or 1503 does not extend radially 
beyond the flywheel 1401 or the housing 1522. 
The improved friction clutch with automatic compensation for wear upon one 
or more parts (e.g., with the adjusting unit 1545 of FIG. 36) can be 
utilized with particular advantage in driving units which are used in 
motor vehicles, especially in vehicles employing at least partially 
automatic (including automatic and semiautomatic) transmissions. The 
friction clutch is then installed between a prime mover (such as the 
engine of a vehicle) and the transmission and is operated or controlled at 
least in dependency upon the operation of the at least partly automatic 
transmission. It is presently preferred to establish a fully automatic 
control for the friction clutch. Automated and fully automatic controls 
for a friction clutch are disclosed, for example, in published German 
patent application Serial No. 40 11 850.9 to which reference may be had, 
if necessary. 
In heretofore known driving units which employ an automatic or 
semiautomatic transmission and a conventional friction clutch, actuation 
of the friction clutch and the design of actuating means (such as electric 
motors and/or cylinder and piston assemblies) present numerous problems. 
Actuation of a conventional friction clutch necessitates the application 
of a relatively large disengaging force which, in turn, necessitates the 
use of rather bulky and powerful actuating means therefor. This 
contributes to the weight, space requirements and cost of such driving 
units, i.e., of units which employ at least partly automated transmissions 
in conjunction with conventional friction clutches. Moreover, the inertia 
of relatively large, bulky and heavy actuators which are employed in 
conventional driving units prolongs their reaction time. If the actuators 
are cylinder and piston units, the application of relatively large forces 
to actuate the friction clutch necessitates the flow of large quantities 
of a hydraulic or pneumatic fluid which also contributes to longer 
reaction times of such actuators. Moreover, it is necessary to employ one 
or more relatively large pumps which are required to supply the cylinder 
and piston units with requisite quantities of a pressurized fluid. 
Attempts to eliminate some drawbacks of the just discussed conventional 
driving units include the utilization of compensating springs which are 
intended to reduce the actuating force necessary to disengage the friction 
clutch and to thus permit the utilization of smaller (more compact) 
actuators. Reference may be had, for example, to published German patent 
application Serial No. 33 09 427. However, since the disengaging force 
varies during the useful life of a conventional friction clutch (the 
required force is relatively small when the friction clutch is new but 
increases with increasing wear upon the friction linings during the life 
span of the friction clutch), a compensating spring can reduce only a 
relatively small fraction of the normally required disengaging force. If 
one takes into consideration all tolerances, it is still necessary to 
provide actuators which must furnish a disengaging force exceeding that 
which is necessary for an unused conventional friction clutch, and this in 
spite of the utilization of compensating springs. On the other hand, a 
driving unit which employs the improved friction clutch with an adjusting 
unit capable of compensating for wear at least upon the friction linings, 
and with a prime mover as well as an automatic or semiautomatic 
transmission, renders it possible to greatly reduce the disengaging force 
well below that which is required for proper operation of conventional 
driving units. Such reduction can take place directly in the friction 
clutch, and the magnitude of the disengaging force remains practically 
unchanged during the entire useful life of the friction clutch. This 
renders it possible to simplify and thus reduce the cost, bulk and 
reaction time of the actuators with attendant savings in space 
requirements and weight of the entire driving unit. Thus, the driving unit 
can be designed to stand relatively small pressures and/or forces. 
Furthermore, this results in a substantial reduction or even complete 
elimination of losses due to friction and/or decreasing resiliency of 
parts in the disengaging means for the improved friction clutch. 
The improved friction clutch and/or the aggregate or assembly employing the 
improved friction clutch is susceptible of numerous additional 
modifications without departing from the spirit of the present invention. 
For example, the features of various described and shown clutches and/or 
aggregates can be used interchangeably or in combination with each other. 
Furthermore, the improved friction clutch and/or the improved aggregate 
and/or a motor vehicle which embodies the improved friction clutch or 
aggregate can also embody numerous additional features which are known per 
se but could further enhance the useful life and/or other desirable 
characteristics of the improved friction clutch and/or aggregate. Still 
further, at least some individual features of the aforedescribed friction 
clutches and/or aggregates embody features which are or could be 
considered to be novel and patentable per se. 
Without further analysis, the foregoing will so fully reveal the gist of 
the present invention that others can, by applying current knowledge, 
readily adapt it for various applications without omitting features that, 
from the standpoint of prior art, fairly constitute essential 
characteristics of the generic and specific aspects of the aforedescribed 
contribution to the art and, therefore, such adaptations should and are 
intended to be comprehended within the meaning and range of equivalence of 
the appended claims.