Patent Publication Number: US-7217204-B2

Title: Continuously variable belt drive system

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
   This is a division of U.S. patent application Ser. No. 09/973,373, filed on Oct. 9, 2001 and issued as U.S. Pat. No. 6,962,543, which is a division of U.S. patent application Ser. No. 09/405,188, filed on Sep. 24, 1999 and issued as U.S. Pat. No. 6,406,390 

   BACKGROUND OF THE INVENTION 
   The present invention concerns a continuously variable transfer drive assembly or transmission mechanism, such as the type suited for use in automotive applications to drive accessory devices. More particularly, the invention relates to a mechanically adjustable belt-type pulley system. 
   Automotive vehicles include a cooling system to dissipate heat developed by the vehicle power plant, such as an internal combustion engine. In a typical automotive vehicle, the lubrication system provides some cooling function as hot lubricant is pumped away from the engine. However, the bulk of the cooling requirements for the automotive vehicle is accomplished by air flowing through the engine compartment and across a radiator. Coolant flowing around the power plant extracts heat from the engine, which heat is subsequently dissipated through the vehicle radiator. 
   In automotive vehicles, the engine compartment is designed to permit flow of ambient air through the compartment and past the radiator. In most vehicles, a cooling fan is provided that increases the flow of air across the radiator. In some vehicle installations, the fan is driven by an electric motor that is independent of the vehicle engine. For smaller passenger cars, the electric motor approach can satisfy the cooling needs for the vehicle. However, unlike passenger cars, heavy trucks cannot use electric motors to drive the cooling fan. For a typical heavy truck, the cooling fan would require up to 50 horsepower to cool the engine, which translates to unreasonably high electrical power requirements. 
   In a typical automotive installation, whether light passenger or heavy truck, the cooling fan is driven by the vehicle engine. In one typical installation shown in  FIG. 1 , an engine  10  provides power through a drive shaft  11  to a transmission  12 . Power to the driven wheels is accomplished through a differential  14 . In addition to providing motive power, the engine  10  is also coupled to a transfer drive assembly  15 . This assembly  15  provides power directly to a cooling fan  16  that is preferably situated adjacent the vehicle radiator  17 . 
   A wide range of technologies is available to transmit power from the engine  10  to the rotating cooling fan  16 . For instance, some transfer drive assemblies  15  are in the nature of on/off clutches. The clutches utilize a friction material to engage the fan when the clutch is actuated. A belt between an output shaft of the engine and the clutch provides rotational input to the clutch in relation to the engine speed. In another drive assembly, a viscous fan drive relies upon the shearing of viscous fluid within a labyrinth between input and output members of the drive. The engagement of the drive is controlled by the amount of fluid allowed into the labyrinth. Viscous drives suffer from many deficiencies. For instance, drives of this type are inherently inefficient because a great amount of energy is lost in heating the viscous fluid. For many viscous drives, this parasitic power loss can be as high as five horsepower. 
   Another difficultly experienced by viscous fluid fan drives is known as “morning sickness.” When the vehicle is started cold, the fluid in the fan drive is more viscous than under normal operating conditions. This higher viscosity causes the drive to turn the cooling fan at full speed, which causes the cooling system to operate at maximum capacity during a time when the vehicle engine needs to be warming up. A further problem with viscous fan drives is that they require a residual speed even when fully disengaged. This residual speed is usually in excess of 400 r.p.m. and is necessary to allow enough fluid circulation within the drive labyrinth for the drive to re-engage on demand. 
   The most prevalent transfer drive systems for a vehicle cooling system rely upon a continuous belt to transfer rotational energy from the vehicle engine to the cooling fan. In the simplest case, one pulley is connected to an output shaft of the engine and another pulley is connected directly to the cooling fan. In this simple case, the speed of the cooling fan is directly tied to the engine, varying only as a function of the fixed diameters of the two pulleys. Typically, the ratio of these diameters generates a speed ratio greater than 1:1—i.e., the fan pulley rotates faster than the engine pulley. 
   One problem exhibited by fixed pulley fan drives is that the fan speed is limited to the fixed ratio relative to the engine input speed. For most vehicles, and particularly most heavy trucks, the maximum cooling air flow requirements occur at the engine peak torque operating condition, which is usually at lower engine speeds. Thus, in order to achieve the proper cooling flow rates, the cooling fan must be sized to provide adequate cooling at the lower engine speeds. The power generated by a fan is related to the cube of its speed. Thus, a fan sized to cool an engine at a lower speed, such as 1200 r.p.m., is grossly oversized at higher engine operating speeds, such as a typical rated speed of 2100 r.p.m. From a cooling standpoint, the significantly greater cooling power provided at higher speeds is not detrimental. However, this over-sizing of the fan equates to wasted power when the engine is not operating at its peak torque condition. For example, a typical 32-inch cooling fan operating at an engine rated speed of 2100 r.p.m., draws approximately 45 horsepower. Of this 45 horsepower, only a fraction, in the range of 10 horsepower, is actually necessary to meet the engines&#39; cooling requirements at this speed. 
   In order to address the varying cooling needs throughout an entire engine operating range, various cooling systems have been developed. For instance, in one type of system, the blades of the fan are rotated to provide variable flow rates. In another application, the shapes of the fan blades themselves are altered to increase or decrease the flow rate at a constant fan rotational speed. 
   One approach to solving the problem of varying cooling needs in an automotive setting has been the continuously variable transmission (CVT) or variable transfer drive assembly. In its most fundamental design, the CVT utilizes a continuous belt having a V-shaped cross section. The belt is configured to engage conical friction surfaces of opposing pulley sheaves. The continuously variable feature of the CVT is accomplished by changing the distance between the sheaves of a particular pulley. As the sheaves are moved apart, the V-shaped belt moves radially inward to a lower radius of rotation or pitch. As the sheaves are moved together, the conical surfaces push the V-shaped belt radially outward so that the belt is riding at a larger diameter. The typical CVT is also sometimes referred to as an infinitely variable transmission in that the V-belt can be situated at an infinite range of radii depending upon the distance between the conical pulley sheaves. 
   Much of the development work with respect CVT&#39;s has been in providing a continuously variable transmission between a vehicle engine and its drive wheels. In a few instances, CVT&#39;s have been applied as an accessory drive. For example, NTN Corporation has developed a rubber belt CVT system that provides a constant accessory drive speed regardless of engine speed. The system using two spring-loaded adjustable pulleys, each having centrifugal weighs that compensate for changes in engine speed. In this system, as the engine speed increases, the centrifugal weights translate radially outward to exert a force on one sheave pushing it toward an opposing sheave. This change in diameter of the sheave maintains a fixed rotational speed, even as the engine speed increases, by altering the ratio of pulley diameters. This fixed speed is used to maintain a constant alternator speed. 
   Ideally, a transfer drive assembly, such as assembly  15  shown in  FIG. 1 , would turn the cooling fan only as fast as is necessary to maintain an optimal engine temperature. Controlling the cooling fan speed conserves power and improves the engine&#39;s overall efficiency. In addition, the transfer drive assembly should have the ability to turn the fan faster at lower engine speeds than at higher engine speeds, because the cooling requirements for the engine are greater during operation at low speed and high torque. 
   Thus far, no accessory drive assemblies are known that are capable of achieving all of these features. Although the continuously variable transmission has been beneficial in operation of cooling fans, the typical CVT cannot accomplish all of these particular factors. 
   SUMMARY OF THE INVENTION 
   The present invention contemplates a continuously variable belt pulley transfer assembly that addresses these prior deficiencies. In one embodiment, the transfer assembly includes a driving pulley assembly and a driven pulley assembly, with a continuous belt transferring rotary motion therebetween. The pulleys are each formed by forward and rear sheaves that define opposing conical surfaces. The drive ratio between the pulleys is determined by the position of the V-shaped belt between the conical surfaces of the sheaves. 
   In one feature of the invention, one pulley assembly, preferably the driving assembly, includes a belt tensioning mechanism that maintains proper belt tension at any speed and pulley drive ratio. The mechanism can include a weight arm that is pivotably mounted to a floating sleeve. The forward and rear sheaves forming the driving pulley are mounted to the floating sleeve for rotation with the sleeve. The sleeve is splined to a rotating drive shaft so the sleeve can slide freely along the drive axis while rotational motion is transmitted to the sleeve. The floating sleeve allows the driving pulley to align itself with the driven pulley when the driven pulley adjusts the drive ratio. 
   Rotation of the floating sleeve causes the weight arm to swing radially outward due to centrifugal effects. The weight arm bears against a roller mounted on the rear sheave, thereby providing an axial force to push the rear sheave toward the relatively stationary forward sheave. As the floating sleeve and driving pulley rotate faster, the axial force generated by centrifugal movement of the weight arm increases. 
   In another aspect of the tensioning mechanism, a spring and lever arm configuration is used to maintain proper belt tension as the drive ratio changes. The mechanism uses a spring plate tending to push the rear sheave toward the forward sheave. When the rear sheave is in its forward-most position, a compression spring associated with the spring plate is only slightly depressed so its axial force is minimal. The present invention contemplates a lever arm disposed between the compression spring and the rear sheave that helps maintain adequate axial force even when the spring is at its minimum compression. The lever arm is pivotably mounted to the floating sleeve and includes a roller at its free end that bears against the rear sheave. The compression springs are retained between the floating sleeve and a spring plate that is free to slide axially relative to the driving pulley. The spring plate includes a roller that contacts a cam edge of the lever arm. Spring force is thus transmitted through the spring plate roller, to the lever arm and eventually to the rear sheave via another roller. The cam edge of the lever arm has a curvature that is calibrated to maintain the necessary axial force at all positions of the rear sheave, including its forward-most position. 
   In yet another feature of the invention, one of the pulleys, again preferably the driving pulley, includes a disengagement mechanism that isolates the belt from the rotation of the pulley. In one embodiment, the disengagement mechanism includes an idler pulley portion between the forward and rear sheaves of the driving pulley. The idler pulley portion defines conical surfaces that transition into the conical surfaces of the primary pulley sheaves. The idler pulley portions are isolated from the forward and rear sheaves by bearings. As the belt sinks lower into the pulley groove it eventually contacts the idler pulley portions. At this point, the belt is no longer in contact with the driving pulley sheaves, so rotation of the driving pulley is not translated to rotation of the belt. 
   The invention also contemplates improvements to a driven pulley member. The driven member includes a ratio adjustment mechanism that utilizes an electric motor and gear arrangement to vary the distance of the rear sheave relative to the forward sheave of the pulley. An actuation screw is provided that can be threaded into and out of a split nut by operation of the electric motor. As the actuation screw is threaded into the split nut, it advances along the axis of the driven pulley assembly. As the screw advances it applies pressure through intermediate components on the rear sheave, pushing it axially toward the forward sheave. Conversely, as the actuation screw is unthreaded from the split nut, the axial pressure on the rear sheave is relieved and the sheave moves away from the forward sheave. 
   The invention further contemplates a fail-safe feature that restores the driven pulley assembly to a predetermined drive ratio in the event of a failure of power to the electric motor. In one aspect, this feature relies upon engagement fingers to hold the separable components of the split nut together to maintain the threaded engagement with the actuation screw. Once the components of the split nut are separated, the internal threads of the nut are disrupted and the threaded engagement with the actuation screw is terminated. In one embodiment, a solenoid holds the engagement fingers in contact with the split nut components. When power to the solenoid is interrupted, the solenoid can no longer hold the engagement fingers in position. A return spring can then push the fingers back, allowing the portions of the split nut to expand apart. 
   In accordance with certain features of the invention, once the split nut is disrupted, the actuation screw is driven forward by operation of a large compression spring. As the actuation screw is propelled forward, it causes the rear sheave to be pushed forward until the sheave reaches a predetermined drive ratio position. 
   It is one object of the invention to provide a continuously variable transfer system that provides mechanical adjustment of the drive ratio of the system. A further object is to provide such a system that maintains sufficient tension in the belt at all speeds and drive ratios. 
   A further object of the invention is accomplished by features that restore the transfer system to a predetermined drive ratio on the occurrence of particular failures. Another object is to provide a transfer system that can achieve a wide range of drive ratios. Yet another object achieved by the invention is to provide means for disengaging the continuous belt from rotation under established conditions. 
   These and other objects, as well as several benefits of the invention can be readily discerned from the following written description of the invention, as illustrated by the accompanying figures. 

   
     DESCRIPTION OF THE FIGURES 
       FIG. 1  is schematic representation of an engine, transmission and cooling system. 
       FIG. 2  is a block representation of one type of transfer drive assembly utilizing a continuous belt and rotating pulley according to a preferred embodiment of the invention. 
       FIG. 3  is an enlarged side cross-sectional view of the driving member of the transfer drive assembly depicted in  FIG. 2 . 
       FIG. 4  is a side cross-sectional view of a forward pulley sheave of the driving member assembly depicted in  FIG. 3 . 
       FIG. 5  is a side cross-sectional view of a rear pulley sheave of the driving member assembly shown in  FIG. 3 . 
       FIG. 6  is an end elevational view of the rear sheave shown in  FIG. 5 . 
       FIG. 7  is an end elevational view of a floating sleeve used in the driving member assembly shown in  FIG. 3 . 
       FIG. 8  is a side cross-sectional view of the floating sheave depicted in  FIG. 7 . 
       FIG. 9  is an end elevational view of a spring-plate used in the driving member assembly shown in  FIG. 3 . 
       FIG. 10  is a side elevational view of the spring-plate shown in  FIG. 9 . 
       FIGS. 11 and 12  are side partial cross-sectional representations of the driving member assembly shown with the pulley sheaves in two orientations. 
       FIG. 13  is a side cross-sectional view of a further embodiment of a driving member assembly for use as part of the transfer drive assembly shown in  FIG. 2 . 
       FIG. 14  is a side cross-sectional view of a driven member assembly for use with the transfer drive assembly shown in  FIG. 2 . 
       FIG. 15  is an end elevational view of the driven member assembly shown in  FIG. 14 . 
       FIG. 16  is an end elevational view of a rear sheave of the driven member assembly shown in  FIG. 14 . 
       FIG. 17  is an end elevational view of bearing pressure plate used in the driven member assembly shown in  FIG. 14 . 
       FIG. 18  is an end cross-sectional view of a support shaft used in the driven member assembly shown in  FIG. 14 . 
       FIG. 19  is an end elevational view of a split nut used with the driven member assembly shown in  FIG. 14 . 
       FIG. 20  is a side elevational view of the split nut shown in  FIG. 19 . 
       FIG. 21  is an end elevational view of a retainer for the split nut for use in the driven member assembly shown in  FIG. 14 . 
       FIG. 22  is a side partial cross-sectional view of an alternative embodiment of a driven member assembly for use with a transfer drive assembly as shown in  FIG. 2 . 
       FIG. 23  is a side cross-sectional view of a further alternative embodiment of a driving member assembly for use in the transfer drive assembly depicted in  FIG. 2 . 
   

   DESCRIPTION OF THE PREFERRED EMBODIMENTS 
   For the purposes of promoting an understanding of the principles of the invention, reference will now be made to the embodiments illustrated in the drawings and specific language will be used to describe the same. It will nevertheless be understood that no limitation of the scope of the invention is thereby intended. The invention includes any alternations and further modifications in the illustrated devices and described methods and further applications of the principles of the invention which would normally occur to one skilled in the art to which the invention relates. 
   The present invention concerns a continuously variable transmission, or transfer drive assembly, particularly suited for driving auxiliary devices in an automotive vehicle. Of course, the principles of the invention can be employed in a variety of applications where continuously or infinitely variable speed ratios are desired. 
   In general terms, the invention provides a driving member assembly that incorporates mechanical tensioning features to maintain proper tension on a V-shaped belt driven by the rotating sheaves of the driving pulley. The driving member assembly also includes a disengagement mechanism operable to isolate the belt from the rotation of the pulley sheaves. In another general aspect of the invention, the continuously variable transfer drive assembly includes a driven member assembly that utilizes mechanical gearing to adjust the relative position between the rotating sheaves of the driven pulley. In addition, the driven member assembly includes a fail-safe mechanism that automatically restores the driven pulleys to a predetermined pitch or pulley ratio upon failure of power supplied to the components of the driven member assembly. 
   With this general background, further details of the various embodiments of the invention will be disclosed with specific reference to the figures. Referring first to  FIG. 2 , the general components of the transfer drive assembly  15  according to one embodiment is shown. In particular, the transfer drive assembly  15  includes a driving member assembly  20  that is connected to a source of rotary power, such as an internal combustion engine, and a driven member assembly  22 , which is connected to a driven device, such as an auxiliary device associated with a vehicle. In the illustrated embodiment, the driven member assembly  22  can be connected to a cooling fan forming part of the engine cooling system. A continuous belt  24  is connected between the pulleys of the driving member assembly  20  and the driven member assembly  22 . The belt  24  is preferably V-shaped and can be of a variety of known configurations and materials. In the preferred embodiment, the belt  24  is driven by frictional contact with the pulley of the driving member assembly. Likewise, the driven member assembly  22  is propelled through frictional contact with the rotating belt. 
   In the present embodiment, the driving member assembly  20  includes a driving shaft  26  that can be configured to mount to the drive shaft of the engine or an auxiliary or PTO shaft driven by the automotive engine. The driven member assembly  22  can include a fan mounting cover  44  with a pattern of screw bores  45  ( FIG. 14 ) to which the engine cooling fan can be engaged. 
   The present invention contemplates a conical pulley system engaged by the continuous belt to transfer rotary power from the driving member assembly  20  to the driven member assembly  22 . Thus, the driving member assembly  20  includes a rear sheave  28 , having a conical engagement surface  29 , and a forward sheave  30 , also having a conical engagement surface  31 . As is well known in the art, the two sheaves  28  and  30  combine to form a pulley for driving the continuous belt  24 . The V-shape of the belt  24  conforms to the opposing conical surfaces  29  and  31  to provide solid frictional contact during rotation of the driving member assembly  20 . 
   The driving member assembly  20  further includes a belt tensioning mechanism  32  that is preferably operably engaged to the rear sheave  28 . The tensioning mechanism maintains tension in the rotating belt  24  by providing pressure to the rear sheave  28 . Pressure on the rear sheave  28  pushes it toward the forward sheave  30  which consequently narrows the gap between the conical surfaces  29 ,  31 . As this gap is narrowed, the continuous belt  24  is urged radially outward to thereby maintain appropriate tension on the belt. 
   For most pulley belt-driven automotive systems, the position of the driving and driven pulleys is fixed to maintain appropriate tension in the belt. However, with the use of a continuously variable system, the belt  24  can be driven by or drive the appropriate pulleys at differing radii. Consequently, the belt tensioning mechanism  32  is important to maintain proper belt tension, ensure efficient transfer of rotary motion between the two pulleys, and eliminate belt squeal associated with a loose or worn belt. 
   In a further feature of the driving member assembly  20 , the pulley formed by the rear sheave  28  and forward sheave  30  is permitted to slide axially along the driving shaft  26 . Changing the pulley ratio between the driving member assembly  20  and driven member assembly  22  causes the centerline of the belt  24  to shift axially relative to the driving shaft  26 . Thus, the pulley formed by the sheaves  28 ,  30  must be free to slide axially to maintain proper alignment between the driving member pulley and driven member pulley. Without this feature, the continuous belt  24  will be skewed between the two pulleys, increasing belt wear and the risk of belt breakage. In the illustrated embodiment, the axial travel of the sheaves is limited at one end by the flange of the driving shaft  26 , and at an opposite end of the driving shaft  26  by a travel stop  34 . 
   A second component of the continuously variable drive assembly  15  is the driven member assembly  22 . The assembly  22  can be fixed to the vehicle, preferably to the engine, by way of a mounting base plate  38 . The driven member assembly  22  also defines a rotating pulley by the combination of a rear sheave  40  and a forward sheave  42 . As with the driving member, the two driven sheaves  40 ,  42  define conical engagement surfaces  41 ,  43 , respectively. A fan mounting cover  44  is engaged to the forward sheave  42  so that rotation of the pulley sheaves causes rotation of the cover  44 , and ultimately rotation of a fan attached to the cover. 
   In accordance with the preferred embodiment of the invention, the continuously variable ratio feature of the assembly  15  is accomplished by a ratio adjustment mechanism  46  integrated into the driven member assembly  22 . In general terms, the adjustment mechanism  46  adjusts the position of the rear sheave  40  relative to the forward sheave  42  to increase or decrease the gap between the two sheaves. As explained above, moving the two sheaves together causes the belt  24  to be forced radially outward to a larger driven radius. Similarly, moving the two sheaves apart allows the belt to drop deeper into the pulley groove, and therefore run at a smaller driven radius. It is preferred that the adjustment mechanism  46  be associated with the driven pulley, rather than the drive pulley. However, a similar mechanism can be incorporated into the driving member assembly  20 , or into both driving and driven assemblies. 
   In a further feature of the preferred embodiment of the invention, the driven member assembly  22  includes a fail-safe mechanism  48 . In one embodiment, the ratio adjustment mechanism  46  is powered by an electric motor. When power is interrupted to the motor, the fail-safe mechanism  48  forces the driven member assembly  22  to a predetermined pulley ratio. Details of the fail-safe mechanism  48  will be developed herein. 
   Referring now to  FIGS. 3–12 , specific features of the driving member assembly  20  will be explained. The driving shaft  26  can include a mounting flange  50  configured to engage a rotating shaft powered by the vehicle engine. The driving shaft  26  defines a splined shaft  51  extending substantially along the length of the driving member assembly  20 . The travel stop  34  in the preferred embodiment can be a snap-ring fixed within a groove at the end of the splined shaft  51 . At the opposite end of the shaft, and adjacent the mounting flange  50 , the driving shaft  26  defines a rear stop surface  52  which further limits the axial travel of the rear and forward sheaves  40 ,  42 . More specifically, the rear stop surface  52  is contacted by a floating sleeve  55  that supports the entire driven member assembly, including the pulley sheaves  40 ,  42 , on the driving shaft  26 . 
   It is understood that the driving shaft  26  and its integral splined shaft  51  are driven by a source of rotary motion. The rotation of the splined shaft  51  is transmitted to the two pulley sheaves through the floating sleeve  55 . The floating sleeve includes inner splines  56  that mate with the splined shaft  51 . This splined interface between the floating sleeve  55  and shaft  51  allows rotary motion to be transmitted between the two components, while permitting the floating sleeve to slide axially along the length of the shaft between the snap-ring  34  and rear stop surface  52 . 
   At an end of the floating sleeve  55  adjacent the travel stop  34 , the sleeve defines outer threads  57 . These threads mate with corresponding inner threads  60  defined in the forward sheave  30 . The outer threads  57  and inner threads  60  are preferably machined threads so that the forward sheave  30  can be firmly engaged, or fixed, to the forward end of the floating sleeve  55 . From the perspective of the floating sleeve  55 , the forward sheave  30  is stationary, meaning that the sheave  30  cannot move axially relative to the sleeve. In contrast, the rear sheave  28  is arranged to slide axially relative to the sleeve  55 . 
   The floating sleeve  55  also defines outer splines  58  situated beneath the rear sheave  40 . The rear sheave  28  then, also defines mating inner splines  62 . Again, the splined interface between the floating sleeve  55  and rear sheave  28  allows the sheave to translate axially along the sleeve, while rotary power is transmitted between the two components. In the preferred embodiment, a collar  63  is disposed around the outside of the rear sheave  28  adjacent the inner spline  62 . In the illustrated embodiment, the rear sheave  28  is movable while the forward sheave  30  is relatively stationary. It is understood, of course, that the roles of the two sheaves of the driving pulley can be reversed, with appropriate modification to the other components of the driving member assembly  20 . 
   In one feature of the invention, the driving member assembly  20  includes a disengagement mechanism  65  at the innermost radius of the pulley formed by the rear sheave  28  and forward sheave  30 . More specifically, the forward sheave  30  defines a bearing recess  61  (see  FIG. 4 ), and the rear sheave  28  defines a similar bearing recess  64  (see  FIG. 5 ). Disposed within the forward bearing recess  61  is a front idler  66  and bearing  68 . The front idler defines a conical surface  67 . Likewise, the rear bearing recess  64  receives a rear idler  69  supported by a rear bearing  71 . The rear idler also defines a conical surface  70  so that the front and rear idlers together define, in essence, a separate conical pulley section. 
   Since the two idlers  66 ,  69  are supported relative to the corresponding sheaves  28 ,  30  by bearings, the pulley formed by the idlers is rotationally isolated from the pulley formed by the sheaves  28 ,  30 . In the operation of the driving member assembly  20 , as the drive assembly  15  moves to a lower ratio, the belt  24  moves lower between the driving member sheaves. When the belt moves far enough, it contacts the conical surfaces  67 ,  70  of the idlers  66 ,  69 , respectively, rather than the surfaces of the primary sheaves  28 ,  30 . When the belt is at this location, the rotation of the belt ceases since the idlers  66 ,  69  do not rotate with the rotating pulley sheaves. In this configuration, the mechanism  65  completely disengages the driven member assembly  22 , and consequently the driven auxiliary device, from the rotary power source. In the case of a cooling fan, when the belt  24  reaches the disengagement mechanism  65 , the rotation of the fan stops. 
   The driving member assembly  20  further includes a belt tensioning mechanism  32 . Since the amount of belt tension required to prevent slip depends on rotational speed, the mechanism  32  applies increasing axial force to the belt as the speed increases. In accordance with a preferred embodiment of the invention, the belt tension is variable instead of constant, to increase the belt life and reduce component fatigue from high belt loads. In other words, at lower rotational speeds, lower belt tension is acceptable. Conversely, at higher speeds, higher belt tension is necessary. Thus, the belt tensioning mechanism  32  is configured to provide greater axial force at higher rotational speeds. 
   The inventive belt tensioning mechanism  32  contemplates two tensioning elements. The first element provides tensioning force as a function of the rotational speed of the driving member assembly  20 . Specifically, this first element is a weight arm assembly  100 . The weight arm assembly  100  includes a number of weight arms  101  that are pivotally mounted to the floating sleeve  55  at a pivot  102 . As shown in more detail in  FIG. 8 , the sleeve  55  defines a weight arm slot  103 , with the pivot  102  at one end of the slot. The weight arm slot  103  provides clearance for pivoting of the weight arm  101 . 
   The weight arm  101  carries a centrifugal weight  104  that is specifically sized to provide a predetermined axial force as a function of rotational speed. In one specific embodiment, the centrifugal weights  104  are formed of depleted uranium due to the high density of the material. In a specific embodiment, the weight arm assembly  100  includes three weight arms  101  symmetrically disposed at 120° intervals around the floating sleeve  55 . At least three weights are preferred to avoid torsional vibration problems. More weight arms and weights can be utilized provided they are symmetrically arranged around the floating sleeve  55 . The magnitude of the centrifugal weights are calibrated based on the maximum required axial force and the centrifugal force generated by rotation of the weights. In the illustrated embodiment where the assembly drives an automotive cooling fan, the weights  104  can be about 1–2 pounds. 
   It is understood that as the floating sleeve  55  rotates with driving shaft  26 , the weight arms  101  gradually pivot outward about pivot point  102  due to centrifugal effects. As the weight arms  101  swing outward, they transmit an axial force to rear sheave  28  to push it closer to the relatively stationary forward sheave  30 . This force transmission occurs through a roller  107 . More particularly, the roller  107  is affixed to the rear sheave  28  through a roller bracket  106 . The bracket is mounted to the rear-most surface of the rear sheave by a mounting screw  108  engaged within screw bore  113  (see  FIG. 6 ). The bracket  106  supports the roller  107  so that as the weight arm  101  presses against the roller, force is transmitted to push the rear sheave  28  axially. 
   The tension in the belt  24  tends to urge the belt deeper into the pulley groove between the sheaves  28 ,  30 . Thus, as the rotational speed of the shaft  26  decreases and the weight arms  101  decline, the belt will act to push the rear sheave  28  rearwardly to maintain constant pressure between weight arm  101  and the roller  107 . In order to further help maintain the weight arm  101  in contact with the roller  107 , a tether in the form of an extension spring  110  is connected between the arm and a spring bracket  109 . The spring bracket is fixed to the rear sheave  28  beneath the roller bracket  106  using the same mounting screw  108 . In the specific embodiment, the spring bracket  109  is partially disposed within a bracket recess  111  (see  FIG. 6 ) to accommodate a reasonable length for the extension spring  110 . The tether or extension spring  110  constantly pulls the weight arm  101  back toward the roller  107 . This prevents problems with the driving member assembly  20  as it initially begins rotating, when the weight arm would ordinarily be fully declined in the absence of any centrifugal effects. Once the shaft  26  starts to rotate, however, the weight arms  101  would be flung outward, which can cause damage to the arms and rollers  107 . The extension spring  110  eliminates this difficulty by keeping the idle position of the arms constrained. 
   Belt tension is not only a function of rotational speed, it is also affected by the drive or pulley ratio—i.e., the ratio between the diameters of the driving and driven pulleys. In order to account for this tensioning relationship, the belt tensioning mechanism  32  includes a second component in the form of a spring pack and lever system. In accordance with one embodiment of the invention, the floating sleeve  55  is configured at its rear end into a number of spring guide blades  75 , shown best in  FIG. 7 . In the illustrated embodiment, three such blades are utilized. Each blade includes two bores through which a spring guide  76  ( FIG. 3 ) extends. An enlarged head  77  of the spring guides  76  prevent their full passage through the blades  75 . A compression spring  80  is mounted over each of the spring guides  76 . In the illustrated embodiment, six such springs are utilized, two each for each guide blade  75 . The compression springs  80  are disposed between the floating sleeve  55  and the rear sheave  28 . Thus, the springs  80  maintain a continuous pressure against the rear sheave  28 , regardless of the position of the belt relative to the pulley sheaves. 
   However, it is well-known that the force supplied by a compression spring is directly related to its displacement. Thus, when the rear sheave  28  is moved to its fullest rearward extent (to the left in  FIG. 3 ), the springs  80  generate their maximum restorative force. By the same token, when the rear sheave  28  is moved to its forward limit of travel, the springs  80  are only minimally depressed, so the force that they apply is considerably weaker. When the belt is at its maximum radially outward position, which can typically correspond with its highest rotational speed, the force being applied by the compression springs  80  is at its lowest, which means that the spring pack is only minimally effective in maintaining tension in the belt  24 . 
   In order to address this problem, a special lever system is incorporated in one feature of the invention. With this feature, a spring plate  82  is slidably disposed over the rear sheave collar  63 . The spring plate defines a spring bore  83 , as depicted best in  FIG. 9 . A spring cup  84  extends though each spring bore  83  and is held in position against the rear surface of the spring plate  82 . The compression spring  80  is then nested within each spring cup  84  so that the springs react against the guide blades  75  of the floating sleeve  55  to push forward against the spring plate  82 . 
   Between each of the spring bores  83  is defined a roller support flange  86 . Each flange  86  supports a spring plate roller  87  engaged at pin bores  87   a . The spring plate  82  further defines a lever slot  88  immediately adjacent or beneath each spring plate roller  87 . The slots  88  are defined to receive a lever arm  90  extending therethrough (see  FIG. 3 ). Each lever arm  90  is pivotally mounted to the floating sleeve  55  at a pivot point  91 . The pivot point is disposed within a lever slot  95  (see  FIGS. 7 and 8 ) so that the lever arm  90  has clearance to pivot relative to the guide blades  75 . The lever arm  90  includes a cam-edge  92  that bears directly against the spring plate roller  87 . The arm further includes a lever arm roller  93  rotatably mounted at the end of the arm opposite the pivot  91 , as best shown in  FIG. 3 . 
   The lever arm roller  93  rides on a force transmitting surface  94  (see  FIGS. 3 ,  5 , and  6 ) defined in the rear surface of the rear sheave  28 . It can thus be appreciated that the force generated by the compression spring  80  and reacted against the guide blades  75 , is applied to the spring plate  82  by way of the spring cups  84 . The spring plate  82  is urged forward (to the right in  FIG. 3 ) so the spring plate roller  87  contacts and pushes the lever arm  90 . As the lever arm  90  is pushed, force is transmitted directly to the rear pulley sheave  28  through the lever arm roller  93 . 
   In the other direction, as the rear sheave  28  moves rearward, or away from the forward sheave  30 , the lever arm  90  rotates about the pivot point  91 . At the same time, the lever arm roller  93  rides radially outwardly along the force transmitting surface  94 . The cam-edge  92  then pushes against the spring plate roller  87  to thereby translate the spring plate actually rearwardly (to the right). As the spring plate is translated, the springs  80  are compressed even further. 
   In a further feature of the driving member assembly, the rear sheave  28  includes a support hub  72 . This support hub underlays the forward sheave  30 . When the rear sheave  28  is at its rearmost position, the support hub  72  is exposed in the gap between the two sheaves, as best seen in  FIG. 12 . 
   This action of the driving member assembly  20  is illustrated in the diagrams of  FIGS. 11 and 12 . In the configuration shown in  FIG. 11 , the driving member assembly  20  is operating substantially at its maximum speed. At this speed, the forward and rear sheaves are united and the support hub  72  is disposed fully underneath the forward sheave  30 . The weight arm  101  is at its greatest radial orientation and the lever arm  90  is at the innermost end of the force transmitting surface  94 . 
   As the speed of the rotational input decreases, the weight arms  101  gradually recline, allowing the rear sheave  28  to translate axially rearward. As the rear sheave moves in that direction, it bears against the lever arm  90  causing the arm to rotate about its pivot point  91 . At the same time, the lever arm, in particular the cam-edge  92 , pushes against the spring plate roller  87 , causing the spring plate  82  to translate axially rearward. This movement compresses the springs  80  (not shown in  FIG. 12 ). 
   In order to maintain a uniform force applied by the compression springs  80 , the cam-edge  92  of the lever arm  90  adopts a predefined curvature. In the specific embodiment, the curvature is a flattened S-shape as shown in  FIG. 3 . This curvature of the cam-edge  92  allows the springs  80  to be pre-compressed to an axial force against the rear sheave  28  sufficient to maintain proper belt tension even at the highest pulley ratios. At the same time, the configuration of the cam-edge  92  regulates the axial force transmitted to the rear sheave  28  as the compression springs  80  are depressed when the driving member assembly  20  is in the configuration shown in  FIG. 12 . 
   In the illustrated embodiment, the spring plate  82  provides a number of spaced openings  89  between each of the roller support flanges  86 . These openings  89  are oriented for passage of each weight arm  101 . As the configuration of the spring plate  82  illustrates, the weight arms are angularly offset from the spring pack portions of the assembly. In the illustrated embodiment, three weight arms are provided, requiring three openings  89  in the spring plate. Of course, additional weight arms can be utilized. It is important, however, to have the arms oriented symmetrically around the driving member assembly to avoid vibration problems associated with an eccentric weight. 
   An alternative embodiment of the driving member assembly is depicted in  FIG. 13 . In particular, the assembly  120  includes a driving shaft  121  having a different configuration for mating with an output shaft of the engine. The assembly  120  includes a rear sheave  123  and a forward sheave  124  that operates similar to the sheaves for the driving member assembly  20 . Both sheaves are supported on a floating sleeve  125  that is actually movable along the length of the shaft  121 . The driving member assembly can also include a disengagement mechanism  126  similar to the mechanism  65  described above. Likewise, the assembly  120  can include a weight arm assembly  127  that centrifugally tightens the belt riding between the sheaves  123 ,  124 . 
   In one modification from the prior embodiment, the floating sleeve  125  supports a spring guide  132  onto which a compression spring  131  is mounted. The rear sheave  123  defines a spring recess  130  in line with the spring guide  132 . The compression spring is then engaged within the recess so that it provides outward forces against the floating sleeve  125  and directly against the rear sheave  123 . In this configuration, the lever arm  90  of the prior embodiment is eliminated. 
   In place of the lever arm, the weight arm assembly  127  includes a specially configured weight arm  133 . Specifically, the weight arm defines a cam-edge  134  that bears against a roller  135  supported on the rear sheave  123 . The cam-edge  134  follows a specific configuration to optimize the axial force applied to the rear sheave  123  at the higher rotational speeds. The cam-edge  134  of the weight arm  133  follows a geometry similar to the cam-edge  92  of the lever arm  90  in the previous embodiment. In both cases, appropriate tensioning force is maintained throughout the range of rotational speeds. 
   Details of the driven member assembly  22  are depicted in  FIGS. 14–20 . As expressed above, the driven member assembly includes a ratio adjustment mechanism  46  that operates on a movable rear sheave  40 . In addition, the driven member assembly includes a fail-safe mechanism  48  that is integrated with the ratio adjustment mechanism  46  to account for a loss of power to the ratio adjustment mechanism. In accordance with a preferred embodiment of the invention, the adjustment mechanism is motor driven. Thus a loss of electrical power to the motor can cause difficulties with respect to the pulley ratio in the absence of a fail-safe mechanism. 
   Turning to  FIG. 14 , it can be seen that the forward sheave  42  is rotatably supported on a needle/thrust bearing  140 . An oil seal  141  is also provided between the rotating sheave and non-rotating components of the driven member assembly  22 . Likewise, the rear sheave  40  is supported on a combination needle/thrust bearing  142 . A rotating seal  143  is also provided between rotating rear sheave  40  and the stationary elements of the driven member assembly. 
   In one feature of the driven member assembly, the rear sheave  40  is interlocked with the forward sheave  42  so that both components rotate together. In order to accomplish the ratio adjustment feature, however, the rear sheave  40  must be permitted to move axially with respect to the relatively stationary forward sheave  42 . Thus, in the illustrated embodiment the forward sheave  42  is provided with a number of slots  144 . The rear sheave  40  includes a like number of interlocking prongs  145 . A preferred arrangement of the slots and prongs is depicted in the end view of the rear sheave  40  shown in  FIG. 16 . It can be seen that the interlocking slots and prongs  144 ,  145  are arc segments. In the specific embodiment, six such interlocking components are provided to adequately transfer torque between the two components and maintain their unison rotational operation. The prongs  145  are configured to readily slide axially along the length of a corresponding slot  144 . 
   The ratio adjustment mechanism  46  relies upon the application of a mechanical force against the rear sheave  40  to move it closer to or further away from the forward sheave  42 . In the preferred embodiment, the adjustment mechanism  46  includes a bearing pressure plate  148  that is at least partially disposed within the rear sheave  40 . The bearing pressure plate  148  directly contacts and presses against the bearing  142  that rotationally supports the rear sheave  40 . The adjustment mechanism  46  further includes a number of force pins  149  that press against the bearing pressure plate  148 . The force pins  149  are supported by a pressure plate  152 . 
   In the preferred embodiment, as shown in  FIG. 17 , the pressure plate  152  includes a plurality of radially extending spokes  153 . A force pin  149  is connected at the end of each of the spokes  153 . Preferably, six such spokes are provided, along with corresponding force pins, uniformly dispersed around the circumference of the pressure plate  152 . In this way, pressure applied by the force pins  149  is evenly distributed against the bearing pressure plate  148 . 
   Movement of the pressure plate  152  is accomplished by operation of an actuation screw  154 . Specifically, the actuation screw  154  includes an enlarged head  155  that bears against the pressure plate  152  through a thrust bearing. The opposite end of the screw  154  defines a screw threaded portion  156 . The threaded portion  156  is configured to threadedly engage internal screw threads  162  of a split nut  158 . In the illustrated embodiment, the split nut is disposed beneath the forward sheave  42 . 
   In operation, the actuation screw  154  is rotated so that the threaded portion  156  is threaded into the split nut  158 . As the actuation screw  154  is continuously threaded, the head  155  bears against the pressure plate  152 , which causes the force pins  149  to push against the bearing pressure plate  148 . Continued rotation of actuation screw  154  ultimately causes the rear sheave  40  to be pushed closer to the forward sheave  42 . As indicated above, moving the two sheaves together pushes their conical surfaces  41  and  43  against the V-shaped belt  24  pushing it radially outward to thereby change the pulley ratio. 
   In order for the actuation screw  154  to accomplish its appointed function, the split nut  158  must be held axially stationary relative to the rear sheave  40 . Thus, the split nut  158  is mounted within a split nut holder  159 . A retainer  160  is internally threaded into the split nut holder  159  to trap the split nut  158  between the holder and the retainer. The split nut holder  159  is itself threaded into a support shaft  164  at a threaded engagement  165 . The support shaft  164  is mounted to the base plate  38 , and is therefore stationary with respect to the ratio adjustment mechanism  46 . 
   Referring to  FIG. 18 , it can be seen that the interior of the support shaft  164  is configured into an array of pin channels  166 . These pin channels are aligned with each of the force pins  149  and with the spokes  153  of the pressure plate  152 . In this way, the pressure plate  152  is prevented from rotating, its movement being limited to axial displacement along the pin channels  166  of the support shaft  164 . 
   As expressed above, the ratio adjustment mechanism  46  is driven by a motor. In the illustrated embodiment, a motor  170  is mounted on the mounting plate  38  by a mounting bracket  169  ( FIG. 15 ). The motor is preferably an electric motor driven by the vehicle electrical system. In a most preferred embodiment, the motor  170  is driven by signals from an engine control module that monitors the engine operation and performance. Specifically, the engine control module can make determinations as to when the transfer drive assembly ratio must be changed and to what extent. Consequently, the motor  170  must be capable of intermittent action and incremental motion. Preferably, the motor  170  is a gear motor driven by a PWM controller, although other motors, such as a stepping motor, can be used. In one specific embodiment, the motor is a model IM-15 motor provided by Globe Motors Co. 
   The motor  170  drives a worm  171  which mates with a worm gear  172 . In the illustrated embodiment, the motor is oriented transverse or perpendicular to the axis B of the driven member assembly  22 . Thus, the worm and worm gear combination transmits the rotary power of the motor to rotational movement of the worm gear  172 . It is understood, however, that other motor and gearing combinations are contemplated by the present invention. For instance, a rack and pinion arrangement can be utilized to translate power from a linear motor to rotational movement. 
   The worm gear  172  is mounted to a worm gear shaft  173 . The worm gear shaft  173  passes through a hollow end of the actuation screw  154 . The worm gear shaft  173  is supported at an opposite end by a thrust bushing  174  mounted within the mounting base plate  38 . 
   The actuation screw  154  defines a pair of opposite engagement slots  175 . A dowel pin  176  passes through the worm gear shaft  173  and is oriented within the engagement slots  175 . In this manner, the worm gear shaft  173  can transmit rotational movement to the actuation screw  154  by way of the dowel pin  176 . At the same time, the actuation screw  154  is free to slide axially along the axis B with the dowel pin  176  sliding along the engagement slots  175 . It can therefore be appreciated that rotation of the worm gear shaft  173  under power from the motor  170  causes direct rotation of the actuation screw  154 . 
   When the motor  170  directs rotation of the worm gear shaft  173  in one direction, the actuation screw  154  is threaded deeper into split nut  158 . As the actuation screw  154  is threaded into the nut it advances toward the rear sheave  40 , pushing the rear sheave as described above. In the alternative, rotation of the motor  170  in the opposite direction causes the actuation screw  154  to unthread from the split nut  158 . As the actuation screw  154  is retracted, the bearing pressure plate  148  moves away from the bearing  142  supporting the rear sheave  40 . The tension within the rotating drive belt  142  causes the belt to project deeper into the gap between the rear and forward sheaves, thereby pushing the rear sheave  40  back toward the pressure plate  152 . Thus, the bearing pressure plate  148  is always substantially in contact with the needle/thrust bearing  142  of the rear sheave  40 . 
   The driven member assembly  22  further includes a fail-safe feature that accounts for a loss of electrical power to the ratio adjustment mechanism  46 . In the preferred embodiment, this mechanism  48  includes a solenoid  180  mounted to the free end of the support shaft  164 . More specifically, the solenoid  180  is supported by a mounting bracket  182  on the split nut holder  159 . A number of control wires  181  electrically connect the solenoid  180  to an external electrical source. Since the support shaft  164  is stationary, the control wires can pass along a channel defined in the shaft, exiting adjacent the mounting base plate  38 . The solenoid  180  is preferably electrically connected to the vehicle electrical system, and most preferably to the engine control module. Thus, when power is interrupted to the adjustment mechanism motor  170 , power is also interrupted to the solenoid  180 . In one specific embodiment, the solenoid  180  can be a low profile push-pull solenoid, such as a model 129415-023 solenoid provided by Lucas Ledex Co. 
   The solenoid  180  includes a solenoid shaft  183  that is held in its actuated position as long as power is provided to the solenoid  180 . The solenoid shaft  183  is threadedly engaged to an engagement finger holder  185 . This finger holder supports a number of engagement fingers  186  that project toward the split nut  158 . More particularly, the engage fingers  186  contact a control ramp surface  161  of the split nut  158 . 
   Operation of the engagement fingers is best understood following an explanation of the structure of the split nut  158 , with specific reference to  FIGS. 19 and 20 . The split nut  158  includes a number of separable components  158   a – 158   c . When the components are combined, they define the internal screw threads  162  that are engaged by the threaded portion  156  of the actuation screw  154 . However, when the components of the split nut are separated, the internal screw threads  162  are interrupted and the threaded portion  156  of the actuation screw  154  has no screw threads to engage. The component  158   a – 158   c  are separated by a split gap  195 . Preferably, this gap is zero when the components of the split nut are combined. On the other hand, when the split nut is separated, this gap  159  is large enough so that the internal threads of the split nut cannot contact the threaded portion  156  of the actuation screw  154 . 
   In order to maintain the integrity of the split nut  158  and insure repeatable separation and combination of its components  158   a – 158   c , the split nut includes a number of guide tabs  196  projecting therefrom. These guide tabs are aligned to slide within corresponding guide slots  197  defined in the retainer  160  (see  FIG. 21 ). The retainer  160  also includes a number of finger bores  198  aligned with the engagement finger holder  185  to receive the engagement fingers  186  therethrough. 
   With this background on the split nut  157 , the operation of the engagement fingers  186  can be more readily understood. As the engagement fingers  186  are pushed rearward, i.e. toward the split nut  158 , the fingers contact the control ramps  161  of each of the split nut components  158   a – 158   c . As the fingers  186  move along the ramp, they continue until they reach the outer diameter of the split nut  158 . At this point, the split gaps  195  are essentially closed and the internal screw threads  162  of the split nut are defined. 
   On the other hand, with the engagement fingers  186  are retracted, they move away from the control ramps  161 . Once the fingers have cleared the ramps and are no longer in contact with the split nut, the components  158   a – 158   c  are free to separate. The overall integrity of the split nut  158  is maintained by the tabs  196  sliding along the slots  197 . The separation of the split nut components  158   a – 158   c  can be accomplished by separation springs  199  mounted within the split nut. The separation springs can be compression springs or leaf springs supported within each component to span the split gaps  195 . 
   During normal operation, the solenoid  180  is powered and the solenoid shaft  183  is maintained in its actuated position. However, when power is removed from the solenoid, the shaft  183  is pushed away from the retainer  160  by operation of a return spring  187 . As shown in  FIG. 14 , the return spring is contained within the engagement finger holder  185  and the retainer  160 . Thus, the return spring  187  in essence pushes the engagement fingers  186  away from the split nut  158 , allowing its components to separate. 
   When the split nut  158  is separated, the threaded portion  156  of the actuation screw  154  no longer has a threaded reaction surface to operate against. In this event, the fail-safe mechanism  48  provides means for pushing the rear sheave  40  forward to the forward sheave  42 , thereby increasing the pulley ratio. This action is accomplished by a return spring  190  disposed within the support shaft  164 . The return spring  190  is situated between a spring carrier  191  at one end and a reaction flange  192  internally formed within the support shaft  164 . The spring carrier  191  is retained relative to the actuation screw  154  by way of a carrier nut  193 . The large return spring  190  can exert force on the spring carrier  191  through a thrust bearing  194  that can be provided to reduce rotational drag on the actuation screw. 
   The fail-safe mechanism  48  of the present invention is operable to return the driven member assembly to a predetermined pulley ratio. For the purposes of explanation, the illustrated embodiment provides a fail-safe ratio of 1:1. When the split nut components  158   a – 158   c  are separated, the response of the fail-safe components depends upon the current pulley ratio. For a ratio greater than the predetermined value (1:1 in the present example), the mechanism  48  drives the rear sheave  40  forward. For ratios less than the predetermined value, the mechanism allows the belt tension to separate the two sheaves. 
   Looking first at a pulley ratio greater than the specific 1:1 value, the rear sheave  40  is separated from the forward sheave  42 . When the split nut components  158   a – 158   c  separate, the threaded portion  156  of the actuation screw is free to slide axially forward along the axis B. The large return spring  190  pushes the spring carrier  191  forward, which contacts the carrier nut  193  to further push the actuation screw  154  forward. As the actuation screw  154  is pushed forward, the enlarged head  155  contacts the pressure plate  152 , causing the force pins  149  to bear against the bearing pressure plate  148 . The bearing pressure plate  148  pushes against the rear sheave  40  until the spring carrier  191  reaches its limit of movement, at which point the rear sheave is immediately adjacent the forward sheave  42 . In a specific embodiment, the two sheaves are separated by a gap of about 0.5 inches at their closest point. 
   The large return spring  190  is calibrated to provide sufficient force to act against the operating tension in the belt  24 . Moreover, the forward movement of the rear sheave is limited by the movement of the spring carrier  191  as the large spring  190  extends. Specifically, in the preferred embodiment, the spring carrier butts against the split nut holder  159  to limit its axial movement. The position of the rear sheave  40  is thus fixed once the carrier contacts the nut holder, which thereby establishes the predetermined pulley ratio. 
   When the pulley ratio is less than the predetermined value (1:1), the spring carrier  191  is already in contact with the nut holder  159 , the threaded portion  156  of the actuation screw  154  extends deeply into the nut holder, and the carrier nut  193  is disposed apart from the spring carrier. When the split nut components  158   a – 158   c  separate, the threaded portion  156  is released and the actuation screw  154  is freely to move axially rearward. The belt tension is then free to push the rear sheave  40  away from the forward sheave  42 . As the rear sheave moves back, the bearing pressure plate  148  pushes against the force pins  149 , which push against the pressure plate  152 , and ultimately against the enlarged head  155  of the actuation screw. As the actuation screw  154  is pushed rearward, the carrier nut  193  moves into contact with the spring carrier  191  which further compresses the large spring  190 . This restorative movement continues until the force generated by the large spring  190  matches the force created by the belt tension. At this point, the driven pulley is at the predetermined ratio. 
   The driven member assembly  22  is indicative of one embodiment of the transfer drive assembly according to the present invention. An additional embodiment is illustrated in  FIG. 22 . Specifically, a driven member assembly  200  includes a rear sheave  201  and forward sheave  202 . In this instance, the fan mounting flange  204  is engaged at one end of a driven shaft  205 . The froward sheave  202  is mounted at the opposite end of the driven shaft  205 . The driven shaft  205  is rotatably supported by a bearing housing  208  by way of a pair of tapered roller bearings  209 . This bearing housing  208  can be mounted to the vehicle or engine. 
   A screw flange  212  is mounted to the bearing housing  208 . The flange  212  defines external screw threads that mate with corresponding threads  215  on a thrust collar  214 . The thrust collar applies force against the forward sheave  201  through a needle bearing  216 . 
   The ratio adjustment mechanism includes a motor  220  that is arranged parallel to the axis of the driven shaft  205 . This configuration for the motor allows the driven member assembly  200  to be mounted within a vehicle having particular space requirements. The motor  220  drives a pinion gear  219  which engages a spur gear  218 . A spur gear  218  is attached to the thrust collar  214 . Thus, rotation of the pinion gear  219  by the motor  220  is translated to rotation of the spur gear  218 . As the spur gear rotates, so does the thrust collar  214 . Rotation of the thrust collar  214  causes its internal threads  215  to advance or retract along the external threads  213  of the screw flange  212 . In this way, the position of the rear sheave  201  relative to the forward sheave  202  can be modified to adjust the pulley drive ratio. 
   In an alternative embodiment of the driving member assembly, an assembly  230  shown in  FIG. 23  includes a driving shaft  232 . The assembly includes a rear sheave  234  and a forward sheave  235 . A disengagement mechanism  236  can be disposed between the two sheaves, as with prior embodiments. 
   The driving member assembly  230  provides a different tensioning mechanism  238  than with the prior embodiments. In particular, the mechanism  238  includes a compression spring  240  that reacts between the driving shaft  232  and a spring cup  241 . A force transfer lever  243  is pivotally mounted at one end to the driving shaft  232 . A transfer roller  244  is provided at the opposite end of the transfer lever  243 . The spring cup  241  includes opposite rollers  246  that rotate along the transfer lever  243 . 
   In operation of this embodiment of the driving member assembly  230 , as the rear sheave  234  moves rearward, it exerts pressure against the transfer roller  244 . This pressure cause the transfer lever  243  to pivot radially outward relative to the driving shaft  232 . As the transfer lever pivots outward, the rollers  246  of the spring cup roll along the lever, causing the spring cup  241  to be displaced axially and rearwardly. As the spring cup moves rearwardly, the compression spring  240  increases its resistant force until equilibrium is established. When viewed in a different sense, the compression spring  240  transfers a tensioning force through the spring cup  241  to the transfer levers  243 , through the rollers  244  and against the rear sheave  234  to push it toward the forward sheave  235 . 
   While the invention has been illustrated and described in detail in the drawings and foregoing description, the same is to be considered as illustrative and not restrictive in character. It should be understood that only the preferred embodiments have been shown and described and that all changes and modifications that come within the spirit of the invention are desired to be protected. For example, in the depicted embodiments, the rear sheave of the drive assembly is movable with respect to the relatively stationary forward sheave. This arrangement can be reversed with appropriate modification to the inventive elements of the system. 
   For instance, in some embodiments, the weight arm assembly, such as assembly  100 , can be mounted differently. In one modification, the weight arms  101  can be pivotably mounted to the rear sheave  40  itself, rather than to the floating sleeve. 
   In alternative embodiments, certain of the features described above can be eliminated. For instance, the disengagement mechanism, such as mechanism  65 , need not be incorporated into all variable ratio transfer drive assembly designs. Likewise, a transfer drive assembly can incorporate several of the aforementioned inventive features, while eliminating the weight arm assembly and/or other components of the tensioning mechanism, such as mechanism  32 . Moreover, other tensioning systems can be substituted for certain specific embodiments.