Patent Document

TECHNICAL FIELD 
     The present invention relates generally to tensioners and more particularly to an asymmetrically damped tensioner utilizing a bearing-ramp plate clutch operatively engaged with the tensioner arm. 
     BACKGROUND 
     It is common for a belt tensioner to have a means to dampen movement of the tensioner arm caused by belt tension fluctuation. The required magnitude of this damping depends on many drive factors including geometry, accessory loads, accessory inertia, engine duty cycle and others. For instance, drive systems that have higher torsional input or certain transient dynamic conditions may require higher damping to sufficiently control tensioner movement. Although higher damping is very effective at controlling arm movement, it can also be detrimental to other critical tensioner functions (e.g. slow or no response to slack belt conditions). In addition, variation or change in damping that occur as a result of manufacturing variation, operating temperature and component break-in or wear can also cause the tensioner to be unresponsive. 
     Timing belt systems have benefited from the use of asymmetric damping to address this problem. An asymmetrically damped tensioner provides damping when additional belt tension is encountered, but is free to respond to slack belt conditions. Although asymmetric functionality may not be required for all other front end accessory drive tensioners, the potential for increased service life, solving other transient dynamic system problems including belt slip during a 1-2 gear shift, or simply making the tensioner less sensitive to damping variation make it a desirable design option. 
     One current solution to this problem uses a viscous linear damper mechanism, such as a shock absorber, attached to a pivoting arm. Asymmetric damping is achieved through, for example, check valves and different orifice sizes in the shock absorber. This solution, however, tends to be expensive and requires more packaging space than a conventional tensioner. Other solutions use wedges that increase damper friction during wind-up or spring loaded self-energizing brake shoe elements. These designs, however, tend to be complex with many small parts to assemble. 
     One-way clutch mechanisms have been proposed, for example in U.S. Pat. Nos. 4,583,962 and 6,422,962, for timing belt tensioners for the purpose of preventing or limiting back travel to prevent tooth jump. These “ratcheting” tensioners, however, lack the ability to relieve belt tension sufficiently when not required. Other timing belt tensioner proposals including, for example, U.S. Pat. Nos. 4,832,665 and 6,375,588, use a one-way device coupled to a viscous damper. Although these devices offer good functionality, retention of the viscous fluid throughout the service life can be difficult. Yet another design disclosed in U.S. Patent App. Publication 2003/0008739, uses friction generated by the clamping action of a wrap spring clutch to provide damping. 
     The aforementioned tensioner designs are not ideal. Accordingly, a new tensioner design is desired. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is an exploded perspective view of one embodiment of a new tensioner with a bearing-ramp plate clutch. 
         FIGS. 2A and 2B  show details of two embodiments of the bearing-ramp plate clutch. 
         FIG. 3  is a view of a partial assembly of the bearing-ramp plate clutch within a tensioner arm. 
         FIG. 4  details the assembled bearing-ramp plate clutch within a tensioner arm. 
         FIG. 5  highlights specific features of one embodiment of a brake plate. 
         FIG. 6  is an exploded perspective view of a new tensioner according to another embodiment with an angled bearing-ramp plate clutch. 
         FIG. 7  is a cut away view of a new tensioner according to another embodiment with an angled bearing-ramp plate clutch. 
     
    
    
     DETAILED DESCRIPTION 
     The new tensioner disclosed herein uses a one-way, bearing-ramp clutch that produces asymmetric friction such that substantial frictional damping is applied to the tensioner only during wind-up (i.e. untensioning). Wind-up results when increasing belt tension causes the belt to lift the tensioner arm in a direction away from the belt. The present invention resists wind-up with a frictional damping force but does not substantially resist movement of the tensioner arm toward the belt with the same frictional damping force. This characteristic is generally known as asymmetric damping. 
     Referring now to the embodiment depicted in  FIG. 1 , the tensioner  21  herein achieves asymmetric damping in part, by the use of a bearing-ramp clutch  7  that connects the tensioner arm  1  to a brake plate  5 . During normal operation, the bearing-ramp clutch  7  is not engaged and the tensioner  21  is biased normally using a spring. During wind-up, the bearing-ramp clutch  7  which is linked to the tensioner arm  1 , expands in size and locks to increase the normal force applied by the bearing-ramp clutch  7  to the brake plate  5 , thereby increasing the frictional damping force and resisting the movement of the tensioner arm  1  away from the belt. Due to the frictional nature of the asymmetric damping, the tensioner  21  will also respond to high belt tension once the frictional force of the damper is overcome, i.e., the tensioner  21  will not restrict all motion in the wind-up direction regardless of the belt tension. 
     The tensioner  21  includes a tensioner arm  1  and a bearing-ramp clutch  7 . The bearing-ramp clutch  7  is comprised of three primary elements, a rotating bearing plate  2 , bearings  3 , and a brake bearing plate  4 . The brake bearing plate  4  is in frictional contact with the brake plate  5 . The interactions between the roller clutch  7  and the brake plate  5  produces the frictional force that generates the asymmetric damping for the tensioner arm  1 . 
     The tensioner arm  1  is pivotable in a first direction A and a second direction B about a pivot axis  15 . As is known in the art, the tensioner arm  1  may be biased by, for example, a torsional spring (not shown) in, for example, the first direction A, to tension an associated power transmitting belt or chain. A pulley (not shown), which is rotatably attached to the distal end  16  of the tensioner arm  1 , is thereby brought into engagement with the belt with a force to tension the belt. The tensioner arm  1  shown in this embodiment also has a cup  11  with a substantially channel shaped cut in the wall of the cup  11  to create a locking slot  12  substantially aligned with the pivot axis  15  of the tensioner arm  1 . 
     In the embodiment depicted in  FIG. 1 , the tensioner  21  further comprises an end cap  6  and tensioner base  8 . The tensioner base  8  is mounted to the opposite side of the tensioner arm  1  from the bearing-ramp plate clutch  7 . The tensioner base  8  has a tensioner axle  9  emerging from the center of the tensioner base  8 . The tensioner axle  9  is substantially parallel the pivot axis  15 . Near the distal end of the tensioner axle  9  there is a series of locking teeth  10  oriented substantially along the pivot axis  15 . The locking teeth  10  are sized to engage the inner locking teeth  14  located around the center of the brake plate  5 . The engagement of the locking teeth  10  with the inner locking teeth  14  substantially prevents the rotation of the brake plate  5  about the pivot axis  15  while substantially allowing the translation of the brake plate  5  along the pivot axis  15 . The end cap  6  is affixed to the distal end of the tensioner axle  9  to cover the end of the tensioner  21  and create an axial force along the pivot axis  15  that compresses the bearing-ramp plate clutch  7  against the brake plate  5 . 
       FIG. 2A  shows one embodiment of a bearing-ramp clutch  7 , comprised of a rotating bearing plate  2 , a brake bearing plate  4 , and bearings  3 .  FIG. 2B  details another embodiment of the rotating bearing plate  2  and the brake bearing plate  4 . In both  FIGS. 2A and 2B , the rotating bearing plate  2  has a locking tab  13 . The locking tab  13  is sized to engage the locking slot  12  on the tensioner arm  1 . When the locking tab  13  is engaged with the locking slot  12 , the rotating bearing plate  2  is able to translate along the pivot axis  15  while substantially linking the rotation of the rotating bearing plate  2  to the rotation of the tensioner arm  1  about the pivot axis  15 . Although the embodiments shown detail a single locking tab  13 , multiple locking tabs may be used to fix the rotation of the rotating bearing plate  2  about the pivot axis  15  to the rotation of the tensioner arm  1 . Alternative structures known to those of ordinary skill in the art can be used to achieve similar functionality including, but not limited to, locking teeth, or mating ovoid or rectilinear profiles. In yet other embodiments, described in more detail below, the locking tab  13  is eliminated and instead frictional forces generated at the interface of the rotating bearing plate  2  and the cup  11  of the tensioner arm  1  urge the rotating bearing plate  2  to follow the rotation of the tensioner arm  1 . 
     The two embodiments shown in  FIGS. 2A and 2B  depict the rotating bearing plate  2  and the brake bearing plate  4  with two different embodiments of the respective bearing raceways  22  and  23 .  FIG. 2A  shows a bearing raceway  22  on the surface of the rotating bearing plate  2  and a mating bearing raceway  23  on the surface of the brake bearing plate  4 . The mating bearing raceway  23  is on the opposite side of the brake bearing plate  4  from the brake surface (not shown in  FIGS. 2A and 2B ). When assembled the bearing raceway  22  and the mating bearing raceway  23  face each other they form a confined space for the roller bearing  3  to travel. Similarly, a second embodiment of the bearing raceway  22  and the mating bearing raceway  23  are shown in  FIG. 2B . The width, and depth of the bearing raceway  22  and the mating bearing raceway  23  vary according to the overall arc length. So for example, in the embodiment shown in  FIG. 2A , the bearing raceway  22  and the mating bearing raceway  23  are longer than the same structures in the second embodiment in  FIG. 2B . The bearing raceway  22  and mating bearing raceway  23  arc length as depicted in  FIG. 2A  is approximately 80 degrees. While, the bearing raceway  22  and mating bearing raceway  23  arc length as depicted in  FIG. 2B  is approximately 35 degrees. In addition to the overall arc length, the taper, and both the depth and width of the bearing raceway  22  and mating bearing raceway  23  relative to the arc length traveled varies. The changing arc length and taper increases the rate of separation between the rotating bearing plate  2  and the brake bearing plate  4  increases as the roller bearing  3  travels within the space formed between the bearing raceway  22  and the mating bearing raceway  23 . 
     The arc length and taper directly effects how much wind-up is necessary for the bearing-ramp clutch  7  to expand and increase the frictional force generated between the brake bearing plate  4  and the brake plate  5  and effectively lock thereby transferring the frictional forces to the tensioner arm  1 . Adjusting the rate of taper of the bearing raceway  22  and the mating bearing raceway  23  thus effectively adjusts the rate of application of asymmetric frictional damping applied to the tensioner  21  during wind-up. The bearing raceway  22  and mating bearing raceway  23  shown in  FIG. 2A  in the first embodiment, has a relatively shallower slope; meaning the rate of change of the position of the bearing  3  changes more slowly than in the case of a shorter arc length bearing raceway  22  and mating bearing raceway  23  shown in the second embodiment in  FIG. 2B . The shorter arc length of the second embodiment, shown in  FIG. 2B , means that for a given displacement of the tensioner arm  1  away from the belt being tensioned, the rotating bearing plate  2  and brake bearing plate  4  will separate further than the first embodiment shown in  FIG. 2A . The resulting increase in separation increases the frictional damping applied to the tensioner  21  to prevent wind-up. 
     The total number of bearings  3  and bearing raceways  22  and mating bearing raceways  23  (collectively, raceways) are determined by the length of the raceways, the taper of the raceways, the size of the bearings  3 . The number of bearings  3  is also dictated by the need for the bearing  3  elements to adequately support and separate the rotating bearing plate  2  and the brake bearing plate  4 . The smallest number of bearing  3  elements for the design is three, and the maximum number is dictated by the size of the bearings  3  and the length and taper of the raceways necessary to achieve a specific damping profile. In the case of the two embodiments shown in  FIGS. 2A and 2B , the total number of raceways and bearing  3  elements is four. The bearings  3  and raceways are equally distributed along the circumference of the rotating bearing plate  2  and the brake bearing plate  4 . Although the embodiments have all of the bearing elements oriented along a single arc line, it is possible to orient the bearings  3  and the raceways in multiple arc lines so they are effectively staggered along the circumference of the bearing-ramp clutch  7 . 
     A partial assembly of the bearing-ramp plate clutch  7  with a tensioner arm  1  is shown in  FIG. 3 . The tensioner arm  1  is shown with the tensioner cup  11  facing upward. A rotating bearing plate  2  is inserted into, and sits in the base of the tensioner cup  11 . The locking tab  13  slides into the locking slot  12  on the interior surface of, the tensioner cup  11 . Four bearings  3  are placed in the bearing raceway  22  in the deepest portion of the bearing raceway  22  or the free running position. The base cap  8  is installed in the opposite side of the tensioner arm  1  and the base cap axle  9  and the base cap locking teeth  10  are protruding from the center of the tensioner cup  11 . The tensioner arm  1  rotates along the pivot axis  15  about the base cap axle  9 . The rotating bearing plate  2  is rotationally coupled to, and rotates with, the tensioner arm  1 . 
       FIG. 4  is an isometric view of another step in an exemplary assembly process whereby the bearing-ramp clutch  7  is now fully assembled inside the tensioner cup  11 . The frictional face or brake surface of the brake bearing plate  4  is shown facing outward, thereby completing the assembly of the bearing-ramp clutch  7 . The opposite side of the brake bearing plate  4  has the mating bearing raceway  23  that is aligned with the bearing raceway  22  with the bearing  3  substantially constrained within. The assembly of the embodiment shown in  FIG. 4  has an assembled bearing-ramp clutch  7 . 
     The brake plate  5  is detailed in  FIG. 5 . The brake plate  5  has a series of inner locking teeth  14 . The inner locking teeth  14  are sized to intermesh with the base cap axle locking teeth  10 . A brake material  50  is located on the periphery of the brake plate  5 . The brake material engages with the mating surface of the brake bearing plate  4  when assembled inside the tensioner arm  1 . The mating frictional interfaces on the brake plate  5  and the brake bearing plate  4  may take many forms other than the flat face to flat face embodiment depicted. Some examples of alternative physical forms for the mating frictional interface between the brake plate  5  and the brake bearing plate  4  include a cup and cone or ball and socket configuration to maximize surface area, discrete brake pads, and other combinations for creating a surface suitable for a friction interface between the elements to provide frictional damping to the tensioner  21 . The frictional surface itself can be fabricated with numerous processes such as heat and surface treatments, surface etching, and processing coupled with material selection in order to control the friction properties of the interface. In the alternative, a dissimilar frictional surface with the desired friction properties is adhered, bonded, glued, welded, or otherwise affixed to the surface. The various techniques for creating effective friction contacts between parts such as these may be selected by those of ordinary skill in the art. 
     Although the particular embodiments shown in the figures depict the use of roller ball bearings, many different types of bearings may be utilized, including, for example, ball, taper, needle, roller, and cylindrical bearings. Additionally, the various components of the bearing-ramp clutch  7  can be fabricated in either fewer or greater numbers of elements. Regardless of the physical structure selected, the bearing-ramp clutch  7  uses a rolling bearing element, bearing  3 , and the interaction between the rolling bearing element inside the bearing-ramp clutch  7  assembly to enable free movement of the tensioner in the tensioning direction and to enable frictional locking during a wind-up condition typified by rotation in the opposite direction. The rolling bearing element is guided by the bearing raceway  22  and the mating bearing raceway  23  and is substantially retained within the bearing-ramp clutch  7  by the rotating bearing plate  2  and the brake bearing plate  4 . 
     An end cap  6  encloses the tensioner  21 , as shown in  FIG. 1 . The end cap  6  encloses the bearing-ramp clutch  7  and provides a compressive force oriented along the pivot axis  15  to the assembly that keeps individual elements of the assembly under compression. The end cap  6  is mounted to the end of the base cap axle  9  and stays substantially fixed relative to the motion of the tensioner arm  1 . The end cap  6  may have an o-ring (not shown) around either or both the external edge and center hole of the end cap  6  to protect the internal elements of the tensioner  21  from dust and dirt and optionally to provide a barrier to prevent the escape of lubricant from inside the tensioner  21 . 
     If the wind-up forces become large enough to overcome the asymmetric frictional damping generated by the interaction between the brake bearing plate  4  and the brake plate  5 , the tensioner arm can still break free and rotate by overcoming the static friction at that interface. In this manner, the tensioner is protected against potential damage to the mechanism caused by extreme wind-up conditions while still providing asymmetric damping suitable to manage normal wind-up experienced during normal engine operation. The selection of the mating friction surfaces on the brake plate  5  and the brake bearing plate  4 , including the bearing-ramp clutch  7  design itself, provide the designer with control over the amount of force necessary to overcome the frictional damping. 
     Through the selection of the mating friction surfaces, the type of bearing  3  used, and the configuration of the bearing raceway  22  and mating bearing raceway  23 , the designer can control the relationship between amount of asymmetric friction applied to the tensioner arm  1  relative to the amount of wind-up experienced. For example, for more aggressive damping, a shorter, more steeply tapering bearing raceway  22  and mating bearing raceway  23  can be used. The shorter, more steeply tapering raceways causes a given change in the position of the tensioner arm  1  to result in a greater displacement of the bearing  3  for a given rotation of the tensioner arm  1  in the direction of wind-up, direction B. The greater displacement moves the rotating bearing plate  2  and the brake bearing plate  4  apart at a faster rate, thereby increasing the normal force applied to the friction surface between the brake bearing plate  4  and the brake plate  5 . The increase in the rate that the normal force is applied to the friction surface for a given change in tensioner arm  1  position, means an increasing rate of asymmetric damping applied to the tensioner arm  1  during wind-up to combat wind-up. 
     A perspective exploded view of a second embodiment of the bearing-ramp clutch  7  in a tensioner  21  is shown in  FIG. 6 . The tensioner cup  11  in the second embodiment is formed with a cup-shaped surface  61 . The rotating bearing plate  2  in the second embodiment has a cone-shaped surface  60 . The cone-shaped surface  60  is sized to mate with the cup-shaped surface  61 . The cone-shaped surface  60  and the cup-shaped surface  61  are friction surfaces that together form a second mating frictional interface between the tensioner arm  1  and the rotating bearing plate  2 . The frictional link causes the rotating bearing plate  2  to move in substantially direction and in substantially the same amount as the tensioner arm  1 . The second embodiment shown in  FIG. 6  utilizes the second mating frictional interface instead of an interlocking slot and tab to rotationally link the rotating bearing plate  2  with the tensioner arm  1 . 
     Inside the rotating bearing plate  2 , bearings  3  are located in contact with bearing raceways  22 . The bearing raceways  22  are fabricated on the interior surface of the rotating bearing plate  2 . The bearings  3  are captured by the brake bearing plate  4 , which has mating bearing raceway  23  (not shown). A torsional spring  63  links the rotating bearing plate  2  with the brake bearing plate  4  such that a rotation of the rotating bearing plate  2  manifests itself as a rotational urge applied to the brake bearing plate  4 . The rotating bearing plate  2  and brake bearing plate  4  with the bearings  3  and the torsion spring  63  form the bearing-ramp clutch  7  assembly of the second embodiment. 
     In the second embodiment, the function of the brake plate  5  is replaced by the end cap  6 . The end cap  6  is fixed to the pivot axis  15  so it cannot rotate. A portion of the surface of the end cap  6 , shown in  FIG. 7 , facing the tensioner arm  1  is a friction surface  70  that interfaces with a friction surface on the brake bearing plate  4 . The mating frictional interfaces on the end cap  6  and the brake bearing plate  4  form a frictional linkage between the bearing-ramp clutch  7  and the fixed end cap  6 . In an alternative embodiment, the brake plate  5 , as described above, is substantially unable to rotate about the base cap axle  9  (not shown in  FIG. 6 ) is used between the bearing-ramp clutch  7  assembly. Regardless of whether or not a brake plate  5  is used in the embodiment, the mating frictional interfaces create a frictional linkage between the bearing-ramp clutch  7  and a substantially fixed surface. 
     Operationally, the tensioner  21  of the second embodiment operates in a similar manner to the first embodiment. When the bearing-ramp clutch  7  is urged due to wind-up of the tensioner arm  21 , (i.e. direction B), the bearings  3  are urged away from the deeper portion of the bearing raceway  22  and mating bearing raceway  23 , the rotating bearing plate  2  and the brake bearing plate  4  are urged apart. When the respective plates ( 2  and  4 ) are urged apart the greater normal force applied to the mating frictional interface and the second mating frictional interface increases the frictional force applied to the damper and thus results in increased asymmetric damping. Similarly, when the tensioner arm  1  is moving in toward the belt (i.e. direction A), the bearings  3  are urged toward the deeper portion of the bearing raceway  22  and the mating bearing raceway  23 . The movement of the bearings  3  into the deeper portion reduces the distance separating the rotating bearing plate  2  and the brake bearing plate  4 . The reduction in distance thus reduces the normal force applied to the mating friction surface and second mating friction surface thereby reducing the friction applied to the tensioner arm  1  as it moves toward the belt (direction A). 
     In the second embodiment shown in  FIGS. 6 and 7 , a torsional spring  63  creates a rotational linkage between the rotating bearing plate  2  and the brake bearing plate  4 . The torsional spring  63  has a pair of tangs  64  at each end of the spring. In the sectional view of the embodiment of the tensioner shown in  FIG. 7 , the tangs  64  of the torsional spring  63  interface with the rotating bearing plate  2  and the brake bearing plate  4  respectively. The torsional spring  63  creates a rotational linkage that couples movement of the rotating bearing plate  2  through the torsional spring to cause a corresponding movement of the brake bearing plate  4 . The rotating bearing plate  2  rotates due to the frictional interface of the cup shaped surface  60  of the rotating bearing plate with the cone-shaped surface  61  on the tensioner arm  1 . 
     Other features depicted in the cut-away of the embodiment shown in  FIG. 7  include the tensioner pulley  73  located at the distal end  16  of the tensioner  1  that applies a force generated by the tensioner  1  to the belt under tension (not shown). The tensioner pulley  73  rotates on tensioner bearings  71  to minimize friction. The tensioner pulley  73  and tensioner bearings  71  rotate about the tensioner axle  72 . 
     Similar to the first embodiment of the tensioner  21 , the second embodiment depicted has friction surface that enable the tensioner arm  1  to tolerate extreme wind-up excursions without damaging the tensioner  21  or its internal components. Specifically, the mating friction surface and second mating friction surface enable the tensioner arm  1  to overcome the asymmetric friction damping and continue to rotate in the event of extreme wind-up it it overcomes the frictional damping created by the bearing-ramp clutch  7 . The second embodiment has a second mating friction surface between the tensioner arm  1 , and the cup-shaped surface  60  inside the cup  11  that interfaces with the rotating bearing plate  2  in addition to the mating friction surface created by the brake bearing plate  4 . The mating friction surface and second mating friction surface can be tailored to work together and selectively give in a way that minimizes the chance to damage to the tensioner  21  during extreme wind-up conditions. 
     The bearing-ramp clutch  7  enables the tensioner  21  to assume two primary operating states, a brake or damping state and a rotate state. As shown in  FIG. 2A , the bearing raceway  22  and the mating bearing raceway  23  provide a deeper portion  26  and a tapering portion  27  of the bearing raceway  22  and mating bearing raceway  23 . The deeper portion  26  is configured to hold the bearing  3  during normal tensioning of the belt. During normal belt tensioning it is desirable to apply a majority of the spring force to the belt under tension with a minimal amount of friction generated by the bearing-ramp clutch  7 . During normal belt tensioning, the spring, not shown in the figures, provides a preload to the tensioner arm  1  thereby urging the tensioner arm  1  toward the belt in direction A. As the tensioner arm  1  moves in direction A, the rotating bearing plate  2  rotates along with the tensioner arm  1 . The rotation of the rotating bearing plate  2  in direction A urges the bearing  3  into the deeper portion  26  of the rotating bearing plate  2  and the brake bearing plate  4 . When the bearing  3  is located in the deeper portion  26 , the rotating bearing plate  2  and the brake bearing plate  4  are the closest together and occupy the least amount of volume inside the space defined along the pivot axis  15  between the end cap  6  and the inside of the tensioner cup  11 . In other words, the distance between the rotating bearing plate  2  and the brake bearing plate  4  is minimized. Since the space occupied by the bearing-ramp clutch  7  is at a minimum, the normal force applied to mating frictional interface is minimized. The brake bearing plate  4  forms the mating frictional interface with either the frictional surface on the brake plate  5 , or the end cap  6 . With the minimal normal force at the frictional contact, the amount of frictional damping generated by the bearing-ramp clutch  7  is at a minimum and it is in a rotate state. 
     The second operating state for the bearing-ramp clutch  7  is the braking state. This condition occurs during tensioner  21  wind-up, when the tensioner arm  1  is pivoting away from the belt being tensioned. As the tensioner arm  1  rotates in direction B, the rotating bearing plate  2  is urged in direction B due to the linkage between the rotating bearing plate  2  and the tensioner arm  1 . The rotation of the rotating bearing plate  2 , as a result of the movement of the tensioner arm  1  from a backlash state, urges the bearings  3  along the bearing raceway  22  and mating bearing raceway  23  away from the deeper portion  26  along the tapering portion  27 . The resulting movement of the bearings  3  into the tapering portion  27 , forces the rotating bearing plate  2  and the brake bearing plate  4  apart from each other. As a result of the movement, overall space occupied by the roller plate clutch  7  increases and the brake bearing plate  4  is urged into the brake plate  5 . Movement along the pivot axis  15  of the brake plate  5  is constrained by the end cap  6  and the movement along the pivot axis  15  of the rotating bearing plate  2  is contained by the tensioner cup  11 . Therefore, the increasing separation of the rotating bearing plate  2  and the brake bearing plate  4  increases the normal force applied to the mating frictional interface formed by the brake bearing plate  4  and either the brake plate  5  or the end cap  6 . The increasing normal force at the mating frictional interface increases the frictional damping. Further, the increasing separation of the rotating bearing plate  2  and the brake bearing plate  4  impedes the movement of the bearing  3  inside the bearing raceway  22  and the mating bearing raceway  23  that effectively locks the rotation of the rotating bearing plate  2  to the brake bearing plate  4 . The effective locking action allows the frictional force generated at the mating frictional interface between the brake bearing plate  4  and the brake plate  5  or the end cap  6  to be transferred to the tensioner arm  1  thereby creating the asymmetric frictional damping needed to resist wind-up of the tensioner  21 . 
     The embodiments described herein include a number of frictional surfaces that are used to create asymmetric damping and effectively brake and/or link the various elements of the tensioner  21 . Regardless of position, the frictional surfaces can take a number of alternative forms within the structure of the overall embodiment including, a cup and cone or ball and socket configuration to maximize surface area, discrete brake pads, and other combinations for creating a surface suitable for a friction interface between the elements to provide frictional damping to the tensioner. The frictional surface itself can be fabricated with numerous processes such as heat and surface treatments, surface etching, and processing coupled with material selection in order to control the friction properties of the interface or a dissimilar frictional surface that is adhered to the surface. The various techniques for creating effective friction contacts between parts such as these may be selected by those of ordinary skill in the art. 
     The embodiments of this invention shown in the drawings and described above are exemplary of numerous embodiments that may be made within the scope of the appended claims. It is contemplated that numerous other configurations of the tensioner assemblies may be created taking advantage of the disclosed approaches. In short, it is the applicant&#39;s intention that the scope of the patent issuing herefrom be limited only by the scope of the appended claims.

Technology Category: 2