GB2392216A

GB2392216A - Belt CVT with failsafe and pulley adjusters.

Info

Publication number: GB2392216A
Application number: GB0325052A
Authority: GB
Inventors: Joshua L Roby
Original assignee: BorgWarner Inc
Current assignee: BorgWarner Inc
Priority date: 1999-09-24
Filing date: 2000-09-11
Publication date: 2004-02-25
Anticipated expiration: 2020-09-11
Also published as: GB2392214A; GB2392216B; GB2392217A; GB0325046D0; GB2392217B; GB2392214B; GB0325056D0; GB2392215B; GB0325052D0; GB2392215A; GB0325050D0

Abstract

A belt continuously variable transmission driven 22 pulley comprises forward sheave 42 and rear sheave 40 axially displacable by motor 170 and worm 171 driven axial screw 154, pressure plate 152, pins 149 and bearing plate 148. The screw reacts against split nut 158 comprising segments spring biassed apart and axially fixed within nut holder 159. In event of motor power failure, solenoid 180 releases and return spring 187 withdraws finger holder 185 and fingers 186 from surrounding the split nut which separates, releasing the screw. If the current ratio is less than a failsafe value e.g. 1:1 belt tension separates the sheaves; if greater spring 190 biasses the sheaves towards each other to achieve failsafe ratio. Driving pulley (20 Fig 3) belt tensioner has sheaves (28,30) biassed together by pivotable centrifugal arm (101) and non-linear spring (80) and cam arm arrangement.

Description

1 2392216

CONTINI JOUSLY VARIABLE BELT DRIVE SYSTEM

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 accGsso--; de-Yiccs. 'Fv,e par+.icJlri., 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 fimction as hot lubricant is pumped awe>' 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, hearty trucks cannot use electric motors to drive the cooling fan. For a typical heavy truck, the cooling fan would require up to 37 low (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 accorr, plished though a differential 14. In addition to providing motive power, the engine 10 is also coupled to a transfer drive assembly 15.

This assembly lS 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 pouter 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 pros ides rotational input to the clutch in relation to the engine speed. In another drive ellJbli, a -viscous fan do ire relies upon the cheering 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 irerent]y inefficient because a great amount of energy is lost in heating the viscous fluid. For mar,y viscous drives, this parasitic power doss can be as high as five horsepower.

Another difficulty experienced by viscous fluid fan drives is known as "morning sickness." When the vehicle is started cold, the fluid in the fan drive is rn ore viscous than under normal operating conditions. This higher viscosity causes the drive to turn the cooling fan at hill 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 fin 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, whicl; is usually at low er 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 naiad speed of 210 r. dill. -Jo.. a cocl1n.=, standpoint, the significantyv 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 81 cm (32-inch) cooling fan operating at an engine rated speed of 2100 r.p.m., draws approximately 33.5 kw (45 horsepower). Of this 33.5 kw (45 horsepower), only a fraction, in the range of 7.4 kw (l0 horsepower), is actually necessary to meet the engines' cooling requirements at this speed.

In order to address the varying cooling needs throughout an entire engine 25 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 (CAT) 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 He 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 s has been in providing a continuously variable transmission between a vehicle engine and. its drive wheels. In a few instances, CVT's have been applied as an accessory drive. For example, NTN Corporation has developed a rubber belt CVT s- yste;= that provides a constant

accessory drive speed regardless of engine speed. The system using two spring-loaded adjustable pulleys each having centrifugal weights 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 cnafg,e ill diml,e.cl of the sheaveain+ains 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 IS shown in FIG. 1, 25 would turn the cooling fan only as fast as is necessary to maintain an optimal engine temperature. Controlling the cooling Plan speed conserves power and improves the engine'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.

The present invention contemplates a continuously variable belt pulley transfer assembly that addresses these prior deficiencies. According to the present invention there is provided a variable ration drive system connectable between a source of rotary motion and a driven device, said system comprising: a driving member connectable to the source of rotary motion for rotation about a drive axis; a driven member connectable to the driven device; and a belt connected between said driving member and said driven member and operable to transmit rotary motion therebetween, wherein at least one of said driving member and said driven member includes: a first sheave and a second sheave movable along said drive axis relative to said first sheave, each of said sheaves having a conical surface configured for frictional engagement filth the belt; and a non- linear spring assembly, separate from said second sheave, disposed between a reaction surface and said second sheave, said spring assembly applying a non-linear force to said second sheave parallel to said drive axis.

Preferably, the non-linear spring assembly comprises a spring and lever al-Lr

configuration which 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 in-verltiun col,tmplates z letter.. disposed he.,een the comn'-ession 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 ann. 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 a development 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 diving 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.

In order that the invention may be well understood, there will now be described an embodiment thereof, given by way of example only, reference being made to the accompanying drawings, in which: 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 pui]ey of the type of the inverition;

FIG. 3 is an enlarged side cross-sectional view of the driving member of the transfer drive assembly of the invention 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; fI5. 5 is a side c,oss-sectio=Aa! 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- sectionaI view of a driven member assembly for use with the transLcr 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 memDcr assembly shown in Fig. 14; FIG. 17 is an end elevational view of bearing pressure plate used in the driven member assembly shown in PIG. 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 slows ill FIG. 14;

1 1 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, and FIG. 23 is a side cross-sectional view of a further alternative embodiment of a driving re:-,-livel assembly for "se,l the trans,f^r die assembly depicted in FT(-T 2.

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 he employed in a variety of applications where continuously or infinitely variable speed ratios are desired. In general teens, the invention provides a driving member assembly that incorporates mechanical tensioning features to maintain proper tension on a 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 venous 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 lS 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 drivels member assembly 2, which is conr,ected 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 1,enlber assembly 20 and the driven member assembly 22. The belt 94 is preferably V-sll.aped and car, 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 drilling member assembly.

Likewise, the driven member assembly 22 is propelled through frictional contact with the rotating belt.

In iDe present, erl,bodi,,.cilt, the d'ivil.g meander assembly 90 inc] ludes 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 4: (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 sheathe 30, also hasting 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 reavowed, 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 squea] associated with a loose or worn belt. In a further feature of the driving melnber assembly 20, tile pulley foisted by

the rear sheave 28 and forward sheave 30 is permitted to slide axially along the drip ing shaft 26. Changing the pulley ratio between the driving member assembly 20 and driven member assembly 22 causes the centerline of the belt 04 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 n-ln.am proper alig,l,ment be+.v.,een the dri!in.= 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 trave] stop 34.

A second component of the continuously variable drive assembly 15 is the 20 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 teens, 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 42 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, flee driven member assembly 22 includes a fail-safe mechanism 48. In one etubc,di;1e:;t, 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 failsafe mechanism 48 ill be developed.

Referring now to FIGS. 3-12, specific features of the driving member assembly 20 wit; be explained. The grieving shaft 6 Cal. include a. ^'unting fl.an,e 50 configured to engage a rotating she* powered by the vehicle engine. The driving shaft 26 defines a splined shaft S1 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 5]. At the opposite end of the shaft, and adjacent the mounting flange SO, the drilling shaft 96 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 40,49, on the driving shaft 26.

It is understood that the driving shaft 26 and its integrral splined shaft 51 are driven by a source of rotary motion. The rotation of the sp] ined 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 hctwccn 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 10 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, wearying that the sheave 28 cannot move axially relative to the sleeve. In contrast, the rear sheave 28 is arranged to slide axially relative to the sleeve 53.

The floating sleeve 5' also defines outer splines 58 situated beneath the rear sheathe 4(). The rear sheave 40, then, also defines notating inner splines 62. Again, the spliced interface between the floating sleeve 55 and rear sheave 28 allows the Cleave

/ 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 sheathe 30 is relatively stationary. It is understood, of course, that the roles of the +.,c sheaves of the drain 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 disengagewcnt mechanism 65 at the innermost radius of the pulley formed. by the rear sheave 28 and forward sheave 30. More specifically, the forward sheave 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 betwocn 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 pnmay 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 in-ven,ion, 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, dower 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.-otatio.la] 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 lOO includes a number of weight arms l OI that are pivotally mounted to the floating sleeve 55 at a pivot 102. As shown in more detai] in FIG. 8, the sleeve 55 defines a weight arm slot 03, with the pivot 102 at one end of the slot. The weight arm slot] 03 provides clearance for pivoting of the weight arm 101.

The weight aim 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 Conned of depleted uranium due to the high density of the-material. In a specific embodiment, the weight am: 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 setrical]y arranged around the floating sleeve 55. The magnitude of the centrifugalweights are calibrated based on the maximum required axial force and the centnfugal 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 0.45 - 0.90 kg (1-2 pounds).

It is understood that as the floating sleeve 53 rotates with driving shaft 26, the weight anus lOI gradually pivot outward about pivot point 109 due to centrifugal effects. As the weight arms lO1 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 ountirig scrc-vv 103 engaged within screen bore

( 113 (see FIG. 6). The bracket 106 supports the roller 107 so that as the weight artn 101 presses against the roller, force is transmitted to push the rear sheave 98 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, Me belt will act to push the rear sheave 29 reanvard1 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 1] O 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 We 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]07. 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] 07. 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 SS is configured at its rear end into a number of spring guide blades 7S, shown best in FIG. 7. In the illustrated embodiment, thTec such blades are utilized. Each blade includes two bores through which a spring guide 76 (Pl03) extends.,1 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:> 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 sheathes.

However, it is well-knon 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 theft maximum restorative force. By the saline o'-en, wher1 the rear sheave 98 is molted 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 mirnmal]y 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 me 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 87a. - 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 am1 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 7i. The lever ann 90 includes a cam-edge 92 that bears directly against the spring plate roller 87. The arm funkier includes a lever ann 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, S. 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 8'2 by play of the spun" cups 84. The spring plate 82 is urged fovard (to the right in FIG. 3) so iLe spring plate roller 87 contacts and

! pushes the lever aim 90. As the lever arm 90 is pushed, force is transmitted directly to the rear pulley sheave 2g through the lever arm roller 93.

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 is, rides radially ouvvard7y along the force transmittin;, 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 20 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 FIG. 11 and 12. In the configuration shown in FIG.]1, 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 predestined 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 sheathe 28 sufficient to maintain proper belt tension even at the highest pulley ratios. At the same fume, the configuration of the cam-edge 92 regulates the axial force transmitted to the rear sheave 28 as the COAmPreSSiOn sprinkles 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 = 101. At the cow;, ration 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 10 supports a spring guide 132 onto which a compression spring 131 is mounted. The rear sheave 123 defines a spring recess 130 inline 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 amp 90 of the prior embodiment is eliminated.

In place of the lever arm, the weight aim assembly 127 includes a specially configured weight arm 133. Specifically, the weight arm defiers 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 193 at the higher rotational speeds. The cam-edge 134 of the weight and 133 follows a geometry similar to the cam-edge 92 of the. lever anal 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 AO. In addition, the Ten 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 tower 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 5 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 interlockirig 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 corfigured 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 fork ard 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 49 that press against the bearing pressure plate 148.

The for ce pills 149 are supported by a pressure plate 159.

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 5 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 159. 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 l 56 is threaded into the split nut] 58. 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]58 between the holder and the retainer. The split nut holder 159 is itself threaded into a support shaft] 64 at a threaded engagement] 65. The support shaft] 64 is mounted to the base plate 3S, and is therefore stationary with respect to the r atio adjustment mecllaiiisin 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] 53 of the pressure plate 152. In this way, the pressure plate 152 is prevented from rotating, its movement being limited to axial dispiacc, llet alongthepin charnels 165 ofthe 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. 1 S). The motor, is preferably an electric motor driven by the vehicle electrical system. In a most preferred embodiment, the motor 1.70 is driven by signals from art 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 TM-IS motor provided by Globe Motors Co (RTM:).

The motor 170 drives a woven 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 oilier motor and gearing combinations are contemplated by the present invention. For instance, a rack and pinion arrangement 25 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 1;4. The worTn gear shaft 173 is supported at an opposite end by a thrust bushing 174 mounted within the mounting base plate 38.

The actuation screw 14 defines a pair of opposite engagement slots 175. A dowe] pin 176 passes through the wonn gear shaft 173 and is oriented within the engagement slots 17. In this manner, the WOnD gear shaft 173 can transmit rotational movement to the actuation screw 154 by way of the dowel pin 175. At the saline time,

! the actuation screw] 54 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.

i.Arhe 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]48 moves away from the bearing 142 supporting the rear sheave 40. The tension within the rotating drive belt l 42 causes the belt to project deeper into the gap between the rear and forward sheathes, 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 farther 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 I 80. In one specific embodiment, the solenoid 180 can be a low profile pushlpull solenoid, such as a model]294 15-

023 solenoid provided by Lucas Ledex Co (RTM).

The solenoid 180 includes a solenoid shaft 183 that is held in its actuated position as long as pouter is provided to the solenoid 180. The solenoid shaft 183 is threadedly engaged to an engagement finger holder 18. This finder holder supports a number of enga;enent fingers 185 that project toward the split nut 158. More

l ( 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 rut 158 includes a number of separable components 158a-15&;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 158a-158c are separated by a split gap 195.

Preferably, this gap is zero when the components of the split nut arc 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 15 Sa-1 58c, the split nut includes a number of guide tabs 196 projecting therefrom. These guide tabs are aligned to slide within corresponding guide slots]97 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] 86 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 158a-158c. 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! 62 of the split nut are defined.

On the other hand, with the engagement fingers 186 are retracted, they move away from the control amps 161. Once the fingers have cleared the ramps and are no longer in contact with the split nut, the components 158a-158c 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 158a-158c can be

I 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 18' is rnaintaine in its actuated position.. However, where Dower is removed from the solenoid, the shaft l 83 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 carTier 191 at one end and a reaction flange 192 internally formed within the support shaft 164. The spring carrier lPl 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 cattier 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. When the split nut components 158a-1S8c 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 rnechanisn allows the belt tension to separate the two sheaves.

Looking first at a pulley ratio greater than the specific 1:1 va]uc, the rear sheave 40 is separated from the forward sheathe 42. When the split nut components 5 8a-158c separate, the tl-eaded portion 156 of the actuation screw is free to slide axially foivad along the axis B. The large return spring 190 pushes the spring cattier

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 lentil the sponge carrier 19! 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 12.7 mm (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 158a-lSSc 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 ] 48 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 enbodinent 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 o-ad sheave 202. In this instance, the fan mounting flan,,e 904 is

- 24 !, engaged at one end of a driven shaft 205. The forward sheave 202 is mounted at the opposite end of the driven shaft 205. The driven shaft 205 is rotatably supported by a beating 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 hearin., housing 9Clk 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 2] 6.

The ratio adjustment mechanism includes a motor 220 that is arranged parallel to the axis of the driven shaft OS. 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 2] 5 to advance or retract along the external threads 213 of the screw flange 212. In this way, the position of the rear sheave 20] 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 duving 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 -20 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 b-ansfer 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 fins embodiment of the driving member assembly 230, as the rear sheave 234 moves rearward, it excepts pressure against the transfer roller 344. This pressure causes the transfer lever 243 to pivot radially ou.v. ard relative to the driving

! f 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 coiiipressio., sp,;.,g 240 transfers a tensiorir,g 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 al] 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 pivotab]y 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.

Cross reference is hereby made to British Patent application no. 0099934. 9

from which the present application is divided.

CLAIMS:

Claims

1. A variable ration drive system connectable between a source of rotary

motion and a driven device, said system comprising: a driving member connectable to the source of rotary motion for rotation about a drive axis; a driven member connecta'ole to the dl.. e,, de; ice; and a belt connected between said driving member and said driven November and operable to trar.smit rotary motion therebetween, wherein at least one of said driving member and said driven member includes: a first sheave and a second sheave movable along said drive axis relative to said first sheave, each of said sheaves having a conical surface configured for frictional engagement with the belt; and a non-linear spring assembly, separate from said second sheave, disposed between a reaction surface and said second sheave, said spring assembly applying a non-linear force to said second sheave parallel to said drive axis (A).

2. A variable ratio drive system according to claim 1, wherein the driving member includes the first sheave, the second sheave and the non- linear spring assembly.

3. A variable ratio drive system according to claim 1 or claim 2, wherein said floating shaft includes a reaction surface; said non-linear spring assembly includes: at least one linear spring bearing at one end on said reaction surface; at least one non- linear force transmission lever disposed between an opposite end of said at least one spring and said second sheave, said lever applying a non- linear force against said second sheave in response to a linear force applied by said spring against said lever; and a spring plate against which said opposite end of said at least one spring bears, said spring plate including a roller mounted thereon to bear against said cam edge of said lever; and said lever Includes: a first end pivotably mounted at said reaction surface;

r ( a roller mounted at an opposite end and arranged to contact said second sheave; and a curved cam edge between said first end and said opposite end against which said spring operates.

4. A variable ratio drive system according to any of the preceding claims, further including an idler element disposed between said first sheave and said second sheave and defining opposite conical surfaces substantially aligned with said first and second conical surfaces and configured for engaging a belt therebetween; and means for supporting said idler element relative to said first and second sheaves to isolate said idler element from rotation with said first and second sheaves.