Abstract:
A dual clutch system having a first and a second friction disk and having a first and a second pressure plate is disclosed. The clutch actuator includes a first fixed ramp and second fixed ramp. Each of the first and the second fixed ramps are disposed to closely cooperate with a first bearing and a second bearing, respectively. A first moveable ramp and a second moveable ramp are disposed to closely cooperate with the first bearing and the second bearing respectively. A first release bearing is adapted to move with a first lever. The first release bearing is actuated by movement of the first moveable ramp. A second release bearing is adapted to move a second lever and the second release bearing is actuated by movement of the second moveable ramp. The first and the second levers are disposed to operatively bias the first and second pressure plates. A cover bearing is disposed to support movement of the first and second moveable ramps and the first and second release bearings. Each of the moveable ramps are actuated by a linkage with a separate motor.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS  
       [0001]     This application claims the benefit of U.S. Provisional Application Ser. No. 60/554,898, filed Mar. 19, 2004; U.S. Provisional Application Ser. No. 60/561,687 filed Apr. 13, 2004; U.S. Provisional Application Ser. No. 60/563,323, filed Apr. 19, 2004, and U.S. Provisional Application Ser. No. 60/563,958, filed Apr. 21, 2004. 
     
    
     STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT  
       [0002]     Not Applicable.  
         [0000]     Appendix  
         [0003]     Not Applicable.  
       BACKGROUND OF THE INVENTION  
       [0004]     1. Field of the Invention  
         [0005]     This invention related generally to linear actuation devices and, more particularly, to an actuator system for a dual clutch.  
         [0006]     2. Related Art  
         [0007]     U.S. Pat. No. 6,012,561 discloses a vehicle transmission having a dual clutch system. The dual clutch system includes first and second flywheels as well as first and second friction disk assemblies and first and second pressure plates for pressing against said first and second friction disk assemblies, respectively. The pressure plates are each operatively engaged by an electromechanical clutch actuator. More particularly, the electromechanical clutch actuator engages a complex cam arrangement to engage one of the pressure plates.  
         [0008]     There remains a need in the art for increased simplicity, durability, and economy in starting clutches and their assembly and operation.  
       SUMMARY OF THE INVENTION  
       [0009]     It is in view of the above problems that the present invention was developed. The invention is an actuator system for starting clutches. The actuator system includes a ball ramp and a motor. The motor controllably rotates the ball ramp by adequate means, which may include, as an example only, a gear train reduction.  
         [0010]     As an example only, the actuator system may be applied for the operation of a dual starting clutch system, in which case the first and second clutches are controlled by varying the axial position of their respective control levers. The ball ramps of the actuator systems are preferably, but not necessarily, nested.  
         [0011]     The actuator system can be used for the actuation of single or dual clutches loaded by diaphragms or loaded by levers, as well as for the actuation of multi-disc clutch packs, either wet or dry. The motor can be either electric or hydraulic.  
         [0012]     In another embodiment of the actuator system, the motor drives the ball ramp through a system of pulleys and prestressed wrap spring coils. The motor has two pulleys which have two distinct diameters, and prestressed bands operatively connecting the two motor pulleys and the ball ramp pulley.  
         [0013]     Further features and advantages of the present invention, as well as the structure and operation of various embodiments of the present invention, are described in detail below with reference to the accompanying drawings.  
     
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0014]     The accompanying drawings, which are incorporated in and form a part of the specification, illustrate the embodiments of the present invention and together with the description, serve to explain the principles of the invention. In the drawings:  
         [0015]      FIG. 1  illustrates schematically a dual starting clutch system  100  controlled by a dual ball ramp system  200  actuated by electric motors;  
         [0016]      FIG. 2A  is a detailed view of the dual ball ramp system  200  illustrated in  FIG. 1 ;  
         [0017]      FIG. 2B  is another detailed view of the dual ball ramp system  200  illustrated in  FIG. 1 ;  
         [0018]      FIG. 3A  is a perspective view of the two non rotating nested ramps of the dual ball ramp system  200 ;  
         [0019]      FIG. 3B  is a perspective partial view of a dual ball ramp system  200 ;  
         [0020]      FIG. 4A  is a perspective view of a generic ball ramp system with spiral tracks;  
         [0021]      FIG. 4B  illustrates a first position of the balls;  
         [0022]      FIG. 4C  illustrates a middle position of the balls;  
         [0023]      FIG. 4D  illustrates a third position of the balls;  
         [0024]      FIG. 5A  is a schematic of a generic ramp system used to define the notations used in  FIGS. 5B  to  5 G;  
         [0025]      FIG. 5B  is a graph illustrating control force versus axial control travel for a clutch loaded by levers;  
         [0026]      FIG. 5C  is a graph illustrating the variation of the control torque consequent to the control force illustrated in  FIG. 5B  versus its rotation for a constant pitch ball ramp system;  
         [0027]      FIG. 5D  is a graph illustrating the constant control torque consequent to the control force illustrated in  FIG. 5B  versus its rotation for a ball ramp system designed with a continuously variable pitch;  
         [0028]      FIG. 5E  is a graph illustrating the angle of rotation of a ball ramp versus its axial travel;  
         [0029]      FIG. 5F  is a schematic illustrating the relative extreme positions of the ramps of a continuously variable ball ramp system designed according to the graph of  FIG. 5E ;  
         [0030]      FIG. 6A  illustrates an alternate embodiment of the actuator system;  
         [0031]      FIG. 6B  illustrates a side view of the actuator system of  FIG. 6A ; and  
         [0032]      FIG. 6C  illustrates a front view of the actuator system of  FIG. 6A  with transverse motor mount.  
     
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS  
       [0033]     The following description describes the application of the actuator system  400  to a dual dry starting clutch of the type loaded by a series of control levers distributed circumferentially, and in which one of the clutches is controlled by pulling on its control levers, while the other clutch is controlled by pushing on its control levers.  
         [0034]     Referring to the accompanying drawings in which like reference numbers indicate like elements,  FIG. 1  illustrates a dual clutch system  100  actuated by a dual actuator system  400  composed of a dual ball ramp system  200  and two motors  111   a  and  111   b.    
         [0035]     The dual clutch system  100  has a cover  105 , a flywheel  104 , a first disc  102   a , a first pressure plate  103   a , a first lever  101   a , a second disc  102   b , a second pressure plate  103   b , and a second lever  101   b . Conventionally, the pressure plates  103   a  and  103   b  are held rotationally relative to the pressure plate  104 , respectively by a series of three circumferential spaced spring straps  113   a  and  113   b . The straps  113   a  and  113   b  apply also a relatively constant axial force which pulls apart the pressure plates  103   a  and  103   b  away from the flywheel  104 .  
         [0036]     In the depicted embodiment, a first ball ramp  300   a  controls the axial position of the first lever  101   a  through a first release bearing  106   a , and a second ball ramp  300   b  controls the axial position of the second lever  101   b  through a second release bearing  106   b . The first clutch is of the pull type lever  101   a , and the second clutch of the push type lever  101   b , with the advantage when combined with a cover bearing  107 , that the preload of the control bearings  106   a  and  106   b  are consequent to the force applied to the levers  101   a  and  101   b  by the straps  113   a  and  113   b  respectively, and such, does not need separate preload springs located between the clutch housing and the control bearing as for conventional starting clutches.  
         [0037]      FIGS. 2A and 2B  illustrates in greater detail the dual ball ramp system  200  illustrated in  FIG. 1 . The dual ball ramp system  200  is composed of the first ball ramp system  300   a  and the second ball ramp system  300   b.    
         [0038]     The first ball ramp system  300   a  is composed of a ramp  224   a  rotatable around the axis of rotation  115  of the dual clutch system  100 , a ramp  223   a  held against rotation relative to the housing of the starting clutch (not illustrated), and one or more balls, one of these being the ball  225   a . In the embodiment illustrated in  FIG. 1 , the ramp  224   a  is rotatably driven through the gears  112   a  and  108   a . Alternatively, the ramp  224   a  is operatively connected to a first electric motor  111   a  by a belt system, or other similar methods. As the ramp  224   a  rotates, the control bearing  106   a  moves axially. The control bearing  106   a  is operatively connected to the first lever  101   a  through a sleeve  221 .  
         [0039]     The second ball ramp system  300   b  is composed of a ramp  224   b  rotatable around the axis of rotation  115  of the dual clutch system  100 , a ramp  223   b  held against rotation relative to the housing of the starting clutch (not illustrated), and one or more balls, one being of these being the ball  225   b . The ramp  224   b  is rotatably driven by the gears  112   b  and  108   b . As the ramp  224   b  rotates, the control bearing  106   b  moves axially. The control bearing  106   b  is operatively connected to the first lever  101   b , and preferably, actuates directly the lever  101   b.    
         [0040]     The non-rotating ramps  223   a  and  223   b  are fastened to a support  109  which is located axially relative to the clutch cover  105  by a cover bearing  107 , and is held against rotation relative to the housing (not shown) of the dual clutch system  100  by adequate means. Alternatively, the cover bearing  107  is removed and the support  109  is fastened by adequate means to the housing of the dual clutch system  100 , in which case the dual ball ramp system  200  is held relative to the housing of the dual clutch system  100  both rotationally and axially.  
         [0041]     The dual ball ramp system  200  is insulated from the rotation of the engine and from the axial vibrations of the engine by the three thrust bearings, i.e., the release bearings  106   a  and  106   b , and the cover bearing  107 .  
         [0042]     The first and second motors  111   a  and  111   b  independently rotate the first and second ramps  224   a  and  224   b , through a preferably a single gear reduction mechanism composed of gears  112   a  and  112   b  driven by the motors  111   a  and  111   b , and driving respectively the gears  108   a  and  108   b . Consequent to said rotation, the ramps  224   a  and  224   b  move axially, thereby acting on the first and second clutch levers  101   a  and  101   b . Movement of the first and second clutch levers  101   a  and  101   b , correspondingly engages or disengages the respective pressure plate  103   a  and  103   b . Accordingly, the engagement and disengagement of the first and second clutch discs  102   a  and  102   b  is controlled by controlling the rotational positions of the first  111   a  and second  111   b  motors.  
         [0043]      FIG. 3A  illustrates how the ramps  223   a  and  223   b  are nested together back to back and the tracks of the balls extend circumferentially as well as radially, and how each is composed of preferably three sections  229   x ,  229   y  and  229   z , each of said sections being fastened by adequate means to the support  109 . Because the ramps  223   a  and  223   b  are nested, the total axial space required for the ball ramp systems  300   a  and  300   b  is substantially reduced.  
         [0044]      FIG. 3B  illustrates that, because the ramps  223   a  and  223   b  are nested, the ramps  224   a  and  224   b  rotate in opposite directions. The ramps  223   a  and  223   b  are fastened to the cover bearing  107  through three helical circumferential segments, the segment  109 ′ being visible in the bottom of  FIG. 2A . The ramps  223   a  and  223   b  are fastened to the support  109  through its three helical sections.  
         [0045]      FIG. 4A  is a perspective view of two generic ramps  410  and  420  facing each other, each having three tracks, respectively  411   a ,  411   b ,  411   c  and  421   a    421   b  and  421   c . Three balls  430   a ,  430   b  and  430   c  roll respectively on the tracks  411   a  and  421   a ,  411   b  and  421   b ,  411   c  and  421   c .  FIGS. 4B  to  4 C illustrate the position of the balls relatively to their tracks when the ramps  410  and  420  rotate relative to each other. As apparent, because the tracks have radially a spiral shape, the balls, balls  430   a ,  430   b  and  430   c  are held automatically circumferentially, and radially, in a same relative position for all relative rotational positions of the ramps  410  and  420 .  
         [0046]      FIG. 5A  is a schematic of a generic ball ramp system having a non rotating ramp  531  and a rotating ramp  532 , and is used for the definition of the various parameters used in  FIGS. 5B  to  5 F. Fc is the axial reaction force applied to the ramp  532  by the control levers, B and x are respectively the angle of rotation and the axial movement of the ramp  532 , T 2  is the external control torque necessary to rotate the ramp  532  consequent to the force Fc. Finally R is the radius of the tracks of the ramps  532  and  531  when assuming that said tracks lay at a constant distance R from the axis of rotation of the ramp  532 .  
         [0047]     The following description of  FIGS. 5A  to  5 F makes reference to the parts of the first ball ramp system  300   a . Said description is identical for the second ball ramp system  300   b.    
         [0048]      FIG. 5B  illustrate the reaction force of the clutch levers (i.e. the control force Fc) as a function of the axial travel (control travel x) of said control levers. For example, for the clutch loaded by the lever  101   a  when the control travel x varies between 0 and 8 mm, the control force Fc may start at around 100 N and reach about 120 N at the kissing point. The kissing point  540  is defined as the point were the pressure plate touches the disc, and is typically reached for a control travel x of 8 mm. Thereafter, while the control travel x varies from 8 to a maximum of 10 mm, the control force Fc rises almost linearly to a maximum of 1,600 N. About 0.8 Joules is stored in the straps  113   a  when the control travel x moves between zero and 8 mm, and about 1.7 Joules is stored in the cushion of the disc  102   a  and the straps  113   a  when the control travel x moves between 8 mm and 10 mm.  
         [0049]      FIG. 5C  illustrates the control torque T 2  as a function of its rotation B for a constant pitch ramp  532  loaded by the control fore Fc illustrated in  FIG. 5B . It should be noted that a constant pitch ramp can be embodied as a screw. The control torque T 2  required to rotate the ramp  532  having a constant pitch is proportional to the control force Fc. With a value for the radius R of the ramp typically found in starting clutches, the control torque T 2  would vary between 290 Nmm and 3800 Nmm, and its variation is proportional to the force Fc illustrated in the graph of  FIG. 5B . Because T 2  is proportional to Fc. As illustrated by the double abcissa of  FIG. 5C , with a constant pitch ramp, the rotation B of the ramp  532  and the axial travel x are strictly proportional, and it is assumed that the pitch is such that the ramp  532  rotates by 240 degrees when said ramp moves axially by 10 mm, which implies a pitch of 15 mm per turn (or 360 degrees).  
         [0050]     When the pitch of the ramp  532  is continuously variable instead of constant, it is possible to design the ramps such that the torque T 2  remains constant when the ramp  532  rotates, in spite of the variation of the control force Fc. In this case, the same amount of energy, i.e. 2.5 Joules, is transferred into the clutch, but the torque T 2  has the lowest possible value, and therefore the rated torque of the motor is also at its minimum. In order to achieve this, the pitch, i.e. the relation between an infinitesimal rotation dB and the correspondent infinitesimal axial movement dx, varies by design continuously along the track. The pitch is therefore continuously variable and is calculated such that, for any given axial position x, the torque T 2  consequent to the force F 2  is constant, in spite of the wide variation of F 2  as illustrated in  FIG. 5B . In this case as shown in  FIG. 5D , for the first part of the control (from clutch open to the kiss point  540 ), the control travel x varies by 8 mm for a rotation B of 76 degrees, and for the second part of the control (between the kiss point  540  and clutch fully closed), the control travel x varies by 2 mm for a rotation B of 164 degrees. In the first part of the control a relatively small rotation of the shaft of the motor  111   a  results in a relatively high travel of the pressure plate  102   a , and in the second part of the control, a large rotation of the shaft of the motor  111   a  results in a relatively low travel of the pressure plate  102   a.    
         [0051]     Comparing the  FIGS. 5C and 5D  it can be observed that the maximum of the torque T 2  when the ball ramp system  300   a  is designed with a constant pitch is about six times higher than for a continuously variable pitch ramp (i.e. 3,800 Nmm versus 600 Nmm), and therefore the maximum torque rating of the motor  111   a  is six times less when the ball ramp system  300   a  is designed with a continuously variable pitch.  
         [0052]     The relation between the control force Fc and the control travel x is approximately linear for the first part of the control, as well as for the second part, and therefore the equations giving the relation between the control force Fc and the control travel x are respectively Fc=a 1 *x+b 1  and Fc=a 2 *x+b 2 .  
         [0053]     The pitch is defined for all values of the rotational position of the ramp  532  directly by the relation between the rotation B and the travel x. For the first and the second part of the control, this relation is as follows:  
       B   =         1   T2     *       [       0.5   *     a   1     *     x   ^   2       +       b   1     *   x       ]     0   8     ⁢           ⁢   and   ⁢           ⁢   B     =       1   T2     *       [       0.5   *     a   2     *     x   ^   2       +       b   2     *   x       ]     8   10             
 
         [0054]     Using the values of  FIG. 5B  to define a 1 , a 2 , b 1  and b 2 , the variation of B as a function of x was calculated according to the previous formulas and is illustrated in  FIG. 5E . It should be noted that the curve has no inflexion point and no discontinuity at x equal to 8 mm, which means that the pitch is continuously variable for all rotational positions B of the ramp  532 .  
         [0055]      FIG. 5F  illustrates an initial position  534  and a final position  534 ′ of two facing tracks  537  and  533  of the ball ramp systems  300   a  and  300   b , as well as an initial and a final positions of a ball rolling on these tracks, respectively  536  and  536 ′. In  FIG. 5F  the tracks are illustrated with the shape defined in the curve of  FIG. 5E .  
         [0056]      FIGS. 6A and 6B  illustrate a dual actuator system  500 , an alternate embodiment of the dual actuator system  400  illustrated in  FIG. 1  of the drawings.  FIG. 6A  illustrates a dual actuator system  500  composed of a dual ball ramp system  600  controlled by two motors  611   a  and  611   b  (only one is illustrated).  
         [0057]     The dual ball ramp system  600  is similar to the dual ball ramp system  200  described in FIGS.  1  to  5 F.  
         [0058]     The actuator system  500  includes two motors  611   a  and  611   b  controlling rotationally a first ball ramp system  700   a  and a second ball ramp system  700   b , both said ramps are coaxial with the axis of rotation  615  of a starting clutch. In  FIGS. 6A and 6B , the motor  611   b  has been removed for clarity.  
         [0059]     The actuator system  500  includes the electric motor  611   a  having two pulleys  656   a  and  657   a  fastened to its shaft  659   a , and such pulleys having respectively a diameter d 1  and a diameter d 2  and a width b. In the depicted embodiment, a first end of a band  654   a  is coiled clockwise on the pulley  656   a , wraps the pulley  653   a  of the ramp  623   a , and its other end is coiled counter clockwise on the pulley  657   a . Alternatively, the band  654   a  wraps the pulley  653   a  for more than one turn, and the wrap angle becomes more than 360 degrees. The two ends of the band  654   a  are fastened by adequate means to the pulleys  656   a  and  657   a , which may include as non limiting examples adhesive, laser spot weld or a rivet. The portion of the band  654   a  which is wrapped around the pulley  653   a  is preferably fastened by adequate means over a relatively short length to said pulley  653   a  by adequate means, which may include as non limiting examples adhesive, laser spot weld or a rivet. The band  654   a  is preferably a very thin band or strip of high strength spring steel, which is pre-stressed such that it will wrap tightly around itself in a circular shape in its free state, and having a thickness h in the order of hundredths of a millimeter. Alternatively, and as an example only, the band  654   a  is weaved, or a composite reinforced by, high strength multifilaments of polymers as a non limiting example, Kevlar or Technora. Because the thickness h is three order of a magnitude lower than the diameters of the pulleys  656   a  and  657   a , and because the shaft  659   a  rotates about ten turns over the control range, the diameters d 1  and d 2  for all practical purposes may be considered approximately constant.  
         [0060]     A compensation spring  652   a  is fastened by adequate means on one of its ends to the housing of the starting clutch (not shown) and, on the other end, to the motor  611   a , such that the compensation spring  652   a  applies a constant force F in the direction illustrated in  FIG. 6B , with the result that the coil  654   a  is permanently tighten with a relatively constant force. The compensation spring  652   a  can be embodied as a spiral spring as illustrated, as a helical torsion spring, or any spring mechanism which supplies a relatively constant force over its range of utilization. As discussed in relation to  FIGS. 5A  to  5 F, the torque T 2  required to rotatably control the ramp system  700   a  is constant over the range of the control, and this translates into a constant torque T 1  on the shaft  659   a . The ratio between the torque T 1  and the torque T 2  is equal to  
         T1   T2     =       d1   +   d2       2   ⁢   D           
 
         [0061]     The forces F/2 applied by the band on each pulley generate opposite torques on the shaft  659   a . However, these torques are not equal and opposite if d 1  and d 2  are different, and as a result, a torque T 0  is applied to the shaft  659   a . The actuator system  500  is designed such that, the torque T 0  resulting from the difference in diameter of the pulleys  656   a  and  656   b  together with the magnitude of the force F developed by the spring  652   a , balances the torque T 1  for all control positions. As a result, discounting the friction losses, the power to actuate the starting clutch is theoretically equal to zero.  
         [0062]     When the shaft  659   a  rotates, the distance W varies, and as a result, energy is transferred back and forth between the compensation spring  652   a  and the shaft  659   a  of the motor  611   a.    
         [0063]     In  FIGS. 6A and 6B  the axis  658   a  of the motor  611   a  and the axis of the dual ball ramp system  615  are parallel. It is advantageous to rotate the motor  611   a  and the compensation spring  652   a  by 90 degrees (not illustrated), such that the axis  658   a  of the motor and the axis  615  of the dual ball ramp system  600  are perpendicular, and the pulleys  656   a  and  657   a  are separated by a distance approximately equal to (D-b). In this case, the coil  655   a  uncoils from the pulleys  656   a  and  657   a  with an angle  661   a  equal to about ninety degrees.  
         [0064]     In view of the foregoing, it will be seen that the several advantages of the invention are achieved and attained.  
         [0065]     The embodiments were chosen and described in order to best explain the principles of the invention and its practical application to thereby enable others skilled in the art to best utilize the invention in various embodiments and with various modifications as are suited to the particular use contemplated.  
         [0066]     As various modifications could be made in the constructions and methods herein described and illustrated without departing from the scope of the invention, it is intended that all matter contained in the foregoing description or shown in the accompanying drawings shall be interpreted as illustrative rather than limiting. Thus, the breadth and scope of the present invention should not be limited by any of the above-described exemplary embodiments, but should be defined only in accordance with the following claims appended hereto and their equivalents.