Abstract:
A utility vehicle drivetrain includes an engine driving a continuously variable transmission (CVT) that drives a transaxle that drives at least one wheel. The continuously variable transmission has a CVT turn ratio defined as an engine output rotary speed into the CVT divided by a CVT output rotary speed into the transaxle. The transaxle has a transaxle turn ratio defined as the CVT output rotary speed divided by a transaxle output rotary speed to the at least one wheel. The transaxle turn ratio can be greater than five times the maximum CVT turn ratio. The transaxle turn ratio can be greater than twenty times the minimum CVT turn ratio. The transaxle turn ratio can be greater than 17.

Description:
[0001]    This application is a continuation of U.S. Ser. No. 11/475,730 filed Jun. 27, 2006, which is a continuation of U.S. Ser. No. 10/616,828, filed Jul. 10, 2003, now abandoned, and claims the benefit of U.S. Provisional Application Ser. No. 60/440,215, filed Jan. 15, 2003. 
     
    
     TECHNICAL FIELD OF THE INVENTION 
       [0002]    The present invention relates to drivetrains for utility vehicles. Particularly, the invention relates to continuously variable transmissions and transaxles for small utility carts. 
       BACKGROUND OF THE INVENTION 
       [0003]    Small utility carts are known such as a John Deere GATOR® utility vehicle. Such vehicles are particularly effective to transport people and cargo short distances at relatively low speeds. Such vehicles are routinely used as golf carts, neighborhood vehicles, work vehicles or plant maintenance vehicles. 
         [0004]    The John Deere GATOR® 4×2 vehicle includes an operator area, a rear mounted engine and a transaxle operatively connected to at least one wheel to drive the vehicle. Between the engine output shaft and the transaxle input shaft is arranged a continuously variable transmission that transfers power from the engine output shaft to the transaxle input shaft. The continuously variable transmission includes a primary clutch in the form of a first split sheave mounted for rotation with the engine output shaft, and a secondary clutch in the form of a second split sheave mounted for rotation with the transaxle input shaft. A drive belt is wrapped around the two sheaves. Both first and second sheaves have V-shaped annular races that are defined by a fixed face and a movable face. The width of each race determines the circumference that the belt wraps around the respective sheave. 
         [0005]    The variable clutch system is speed and load sensitive. The primary and secondary clutches work together, automatically up-shifting and down-shifting. The shifting changes the ratio between the clutches allowing the engine to operate at optimum efficiently, at the peak of its power curve. 
         [0006]    The primary clutch is engine speed sensitive, and is mounted on the engine crankshaft. It operates on the principle of centrifugal force. The secondary clutch, mounted on the transaxle input shaft, is load sensitive to the rear drive wheels. 
         [0007]    The primary clutch spins with the engine crankshaft, and centrifugal force on cam weights within the primary clutch tends to close the movable and stationary sheave faces together, while a primary clutch coil spring urges the sheave faces apart. At idle speed the centrifugal force is not enough to overcome force from the spring. The primary clutch split sheave remains opened wide and does not engage the drive belt. 
         [0008]    At a minimum load, the primary clutch sheave faces are moved closer together, and start to move the drive belt. The drive belt wraps a maximum race circumference of the secondary clutch. A high CVT turn ratio between the clutches exists, similar to a low gear operation, as long as there is minimal load. The CVT turn ratio is the ratio of the number of turns of the engine output shaft that turns the primary clutch to the number of turns of the transaxle input shaft that is turned by the secondary clutch. 
         [0009]    As engine speed increases, centrifugal forces of cam weights force the primary clutch to “up-shift”, moving the sheave faces together and forcing the drive belt to an outer race circumference. The belt overcomes force from a secondary clutch spring that is arranged to urge the movable and stationary sheave faces of the secondary clutch together, wherein the drive belt is pulled deep in the secondary clutch, wrapping a minimum race circumference, resulting in a low CVT ratio, similar to a high gear operation. 
         [0010]    Down-shifting occurs as a load is encountered, such as a hill or soft terrain. Turning of a stationary sheave face of the secondary clutch is resisted by the load on the wheels via the transaxle, and at the same time, torque from the drive belt moves a moveable sheave face of the secondary clutch up a ramp that is fixed to turn with the stationary sheave face. The ramp and spring force the movable and stationary sheave faces together which forces the belt to wrap an increased circumference of the secondary clutch. The belt overcomes centrifugal forces of the primary clutch, thus wrapping a decreased circumference of the primary clutch, causing the down-shifting. 
         [0011]    The vehicle drivetrain is operable over a total gear ratio that is defined by: 
         [0000]      CVT turn ratio×transaxle turn ratio=total gear ratio. 
         [0000]    The CVT turn ratio is the ratio of the number of turns of the engine output shaft to the number of turns of the transaxle input shaft. Stated in another way, the CVT turn ratio is the ratio of the rotary speed of the engine output shaft to the rotary speed of the transaxle input shaft. The transaxle turn ratio is typically a fixed ratio of the number of turns of the transaxle input shaft divided by the number of turns of the transaxle output shaft. Stated in another way, the transaxle turn ratio is the ratio of the rotary speed of the transaxle input shaft to the rotary speed of the transaxle output shaft. 
         [0012]    In the heretofore known John Deere GATOR® 4×2, the transaxle turn ratio is fixed at 15.28 and the CVT turn ratio is continuously variable with a maximum CVT turn ratio of 4.2 and a minimum CVT turn ratio of 0.94. The transaxle turn ratio is about 3.6 times the maximum CVT turn ratio. The transaxle turn ratio is about 16.2 times the minimum CVT turn ratio. 
       SUMMARY OF THE INVENTION 
       [0013]    The invention provides a cart-type utility vehicle that incorporates a combination of a continuously variable transmission and a transaxle having a preselected total gear ratio that produces an exemplary tractive force or pulling force while also allowing an acceptable vehicle top speed especially for a relatively lightweight, economically manufactured vehicle. 
         [0014]    The invention provides a cart-type utility vehicle having a drivetrain including an engine driving a continuously variable transmission (CVT) that drives a transaxle that drives at least one wheel. The continuously variable transmission has a CVT turn ratio defined as an engine output rotary speed into the CVT divided by a CVT output rotary speed into the transaxle. The transaxle has a transaxle turn ratio defined as the CVT output rotary speed divided by a transaxle output rotary speed to the at least one wheel. 
         [0015]    According to an exemplary embodiment of the invention, the transaxle turn ratio is considerably greater than heretofore known comparable utility vehicles. The transaxle turn ratio can be greater than five times the maximum CVT turn ratio. The transaxle turn ratio can be greater than twenty times the minimum CVT turn ratio. The transaxle turn ratio can be greater than 17. 
         [0016]    According to an exemplary embodiment of the invention the CVT turn ratio is variable from between about 3 to about 0.8 and the transaxle turn ratio is about 18. 
         [0017]    A total gear ratio is the CVT turn ratio of the continuously variable transmission multiplied by the turn ratio of the transaxle. According to an exemplary embodiment of the invention, the difference between the maximum total gear ratio and the minimum total gear ratio is about 40 or greater. According to an exemplary embodiment of the invention, a maximum torque produced by the vehicle drivetrain corresponds to a total gear ratio of about 57, an engine rotary speed of about 2517, and an axle rotary speed of about 43. 
         [0018]    A particular exemplary embodiment of the invention provides a continuously variable transmission having a CVT turn ratio of about 0.77 to 3.1 and a transaxle with a transaxle turn ratio of 18.35. These ratios, when multiplied, result in a maximum total gear ratio of about 57 and a minimum total gear ratio of about 14. The difference between the maximum total gear ratio and the minimum total gear ratio is about 43. The transaxle turn ratio is about 6 times the maximum CVT turn ratio. The transaxle turn ratio is about 24 times the minimum CVT turn ratio. 
         [0019]    Numerous other advantages and features of the present invention will be become readily apparent from the following detailed description of the invention and the embodiments thereof, from the claims and from the accompanying drawings. 
     
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         [0020]      FIG. 1  is a side view of a utility vehicle incorporating the present invention; 
           [0021]      FIG. 2  is a side view of a powertrain that drives the vehicle of  FIG. 1 ; 
           [0022]      FIG. 3  is a schematic top view of a continuously variable transmission according to the invention in a maximum torque mode; 
           [0023]      FIG. 4  is a schematic side view of the transmission of  FIG. 3 ; 
           [0024]      FIG. 5  is schematic top view of the continuously variable transmission of  FIG. 3  shown in a maximum speed mode; 
           [0025]      FIG. 6  is a schematic side view of the transmission of  FIG. 5 ; 
           [0026]      FIG. 7A  is a fragmentary top half sectional view of an engine driven split sheave of the transmission taken generally along line  7 A- 7 A of  FIG. 3 ; 
           [0027]      FIG. 7B  is a fragmentary bottom half sectional view of the engine driven split sheave of the transmission taken generally along line  7 B- 7 B of  FIG. 5 ; 
           [0028]      FIG. 8A  is a fragmentary top half sectional view of a transaxle-driving split sheave of the transmission taken generally along line  8 A- 8 A of  FIG. 3 ; 
           [0029]      FIG. 8B  is a fragmentary bottom half sectional view of a transaxle-driving split sheave of the transmission taken generally along line  8 B- 8 B of  FIG. 5 ; 
           [0030]      FIG. 8C  is a schematical fragmentary plan view of a portion of the transaxle-driving split sheave shown in  FIG. 8B ; and 
           [0031]      FIG. 9  is a schematic sectional view of a transaxle. 
       
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
       [0032]    While this invention is susceptible of embodiment in many different forms, there are shown in the drawings, and will be described herein in detail, specific embodiments thereof with the understanding that the present disclosure is to be considered as an exemplification of the principles of the invention and is not intended to limit the invention to the specific embodiments illustrated. 
         [0033]      FIG. 1  illustrates a utility vehicle  10  that incorporates the present invention. The vehicle  10  includes a frame  12  carried by front wheels  14  and rear wheels  16 . The vehicle  10  includes a driver&#39;s station  22  and a cargo area  26 . 
         [0034]      FIG. 2  illustrates a drive train  32  for the vehicle  10  that is partially hidden in  FIG. 1 . The drive train includes a directional shifter  34  and a differential lock lever  36 . 
         [0035]    An accelerator pedal  42  is located in the footwell of the vehicle. An engine  52  is mounted in front of a transaxle  56  that drives a rear axle  60  operatively connected to the rear wheels  16 . 
         [0036]    An engine output shaft  64  is fixed to an engine driven primary clutch in the form of a split sheave  66 . A transaxle-driving secondary clutch in the form of a split sheave  70  is fixed to a transaxle input shaft  72 . 
         [0037]    The sheaves  66 ,  70  each provide a variable depth belt race. A belt  80  encircles the shafts  64 ,  72  within the races of the sheaves  66 ,  70 . 
         [0038]      FIG. 3  illustrates the sheave  66  including an actuator  86  that carries a movable plate  88  having a movable face  89 , and a fixed plate  90  that includes a fixed face  92 . A V-shaped sheave race  94  is defined between the faces  89 ,  92 . 
         [0039]    The transaxle-driving sheave  70  includes an actuator  106  that carries a movable plate  110  having a movable face  112 , and a fixed plate  114  having a fixed face  116 . A V-shaped sheave race  120  is defined between the faces  112 ,  116 . 
         [0040]      FIG. 4  illustrates the transmission in a maximum torque mode wherein a radius r 1  of the race  94  is a minimum and a radius r 2  of the race  120  is maximum. According to an exemplary embodiment of the invention r 2 /r 1 =3.1. 
         [0041]      FIGS. 5 and 6  illustrate the transmission in a maximum speed mode wherein a radius of the race  94  is r 3  and a radius of the race  120  is r 4 . According to an exemplary embodiment of the invention r 4 /r 3 =0.77. 
         [0042]      FIGS. 7A and 7B  illustrate the primary clutch  66  to be mounted on the engine output shaft  64 . The centerline thereof is indicated as CL. An internal bore  66   a  is configured to be fixedly coupled to the shaft  64  to rotate therewith. The clutch  66  includes a housing  130 . The fixed sheave plate  92  includes a spindle  134  that enters the housing  130  and is fixed to a backing plate  138  within the housing  130 . The movable sheave plate  88  includes a plurality of cam weights  140  that are pivotally connected to a backside of the movable sheave plate  88  and have the cam surfaces  140   a  that are pressed against pins  144  which are mounted to the backing plate  138 . A coil spring  150  is located surrounding the spindle  134  and between a shoulder  151  of the spindle  134  and a shoulder  152  of the movable sheave plate  88 . The spring  150  urges the movable sheave plate face  89  away from the fixed sheave plate face  92 . 
         [0043]    The clutch  66  operates on the principle of centrifugal force and is engine speed sensitive. At idle speed, the primary clutch  66  spins with the engine output shaft  64 , but centrifugal force on the weights  140  is not enough to overcome the force of primary clutch spring  151 . The primary clutch sheave remains opened wide and does not engage the drive belt  80 . 
         [0044]    As shown in  FIG. 7A , at a minimum load, the primary clutch sheave plate faces  89 ,  92  are moved closer together, by centrifugal force of the cam weights  140  against the pins  144 . The sheave plates  88 ,  90  start to circulate the drive belt  80  at a minimum wrapped race circumference of the clutch  66 . The drive belt  80  wraps a maximum race circumference of the secondary clutch. A high ratio between the clutches exists, similar to a low gear operation, as long as there is minimal load. 
         [0045]    As shown in  FIG. 7B , as engine speed increases, centrifugal forces of the cam weights  140  force the primary clutch to “up-shift”, moving the drive belt to an increasing race circumference. The belt overcomes force from a secondary clutch spring  174  (described below), wherein the drive belt  80  is pulled deep in the secondary clutch  70 , wrapping a decreasing race circumference and giving a low ratio, similar to a high gear operation. 
         [0046]      FIGS. 8A and 8B  illustrate the secondary clutch  70 . The centerline thereof is indicated as CL. An internal bore  70   a  is configured to be fixedly coupled to the transaxle input shaft  72  to rotate therewith. The stationary clutch includes a spindle  170  connected to the fixed sheave plate  114 . A spindle  170  is connected to a backing plate  172 . A coil spring  174  is arranged surrounding the spindle  170  between the backing plate  172  and the movable sheave plate  110 . The spring  174  urges the movable sheave plate face  112  toward the stationary sheave plate face  116 . The backing plate  172  includes one or more ramps  180  and the movable sheave plate includes one or more protrusions  182  which ride on the ramp(s)  180  (see  FIG. 8C ). The protrusion(s)  182  can be in the form of plastic replaceable buttons. 
         [0047]    The secondary clutch  70  is load sensitive to the rear drive wheels  16 . Down-shifting occurs as a load is encountered, such as a hill or soft terrain. The load on the wheels is transmitted to the stationary sheave plate  114  of the secondary clutch through the transaxle, and at the same time, torque from the drive belt  80  moves the moveable sheave plate  110  of the secondary clutch up the ramp  180 . The ramp  180  and spring  174  forces the faces  112 ,  116  closer together and the belt  80  to wrap an increased circumference of the secondary clutch ( FIG. 8A ). The secondary clutch  70  overcomes centrifugal forces of the primary clutch cam weights  140 , thus causing a wrapping of a decreased circumference of the primary clutch, which causes the down-shifting. 
         [0048]    Primary and secondary clutches of the type described above are commercially available from suppliers such as Hoffco-Comet Industries, Incorporated of Richmond, Ind., U.S.A as models 72C and 88D. Examples of primary and secondary clutches are disclosed in U.S. Pat. Nos. 5,647,810; 5,597,060 and 5,967,286, herein incorporated by reference. 
         [0049]      FIG. 9  illustrates, in schematic fashion, a typical transaxle  56 . The transaxle  56  includes the input shaft  72 , a forward drive gear  230  connected thereto, and a forward driven gear  232  that is enmesh with the forward drive gear  230  and fixed to a reduction gear shaft  236 . A reduction drive gear  238  is fixed on the gear shaft  236  and enmesh with a differential gear  244 . The differential gear  244  drives differential pinion gears  248 . The pinion gears  248  drive left and right output shafts  254 ,  256 , that drive the wheels  16 . 
         [0050]    A reverse drive sprocket  260  and a reverse driven sprocket  264  are coupled by a reverse drive chain  266  and are provided for driving the transaxle in reverse. A shift fork  274  and an input shaft  276  are operatively connected to the directional shifter  34  and are used to select between forward and reverse operation. 
         [0051]    The transaxle  56  also includes a neutral switch  292 , brake disks and plates  294 , a brake shaft  296 , brake levers  298 , a brake actuator plate  300 , a ball and ramp arrangement  302 , a differential lock pin  306 , and a differential lock collar  307  all arranged and configured in a conventional manner. 
         [0052]    According to an exemplary embodiment of the invention, the size and number of teeth of the gears  230 ,  232 ,  238 ,  244 ,  248  are selected such that the number of turns of the input shaft  72  is about 18, in particular 18.35, times the number of turns of the output shafts  254  or  256 . A transaxle so configured may be commercially available from Kanzaki Kokyukoki Mfg. Co. Ltd. Of Amagasaki, Hyogo, Japan, as Model AM 131715 or from Transaxle Manufacturing of America Corporation, Rockhill, S.C., as Model AM132333. An alternate transaxle having a transaxle turn ratio of 19.2 to 1 can be obtained from Dana Corporation, Off Highway Systems (Components) Group, Maumee, Ohio, U.S., Model H12. 
         [0053]    An exemplary embodiment utility vehicle of the invention has the following specifications: 
         [0054]    Vehicle weight 650 lbs. 
         [0055]    Maximum payload 800 lbs. 
         [0056]    Rolling radius (20 inch tires) 0.75 ft 
         [0057]    Gross weight with no cargo or operator 1450 lbs. 
         [0058]    Maximum Torque Operation 
         [0059]    Engine RPM 2517 
         [0060]    Engine torque 13.73 ft-lbs. 
         [0061]    Primary clutch pitch diameter 2.64 in. 
         [0062]    Secondary clutch pitch diameter 8.3 in. 
         [0063]    CVT ratio 3.14 
         [0064]    Transaxle (T/A) ratio 18.35 
         [0065]    Total ratio CVT ratio×T/A ratio 57.6 
         [0066]    Input shaft torque 43.17 ft-lbs. 
         [0067]    Axle torque 791 ft-lbs. 
         [0068]    Axle speed 43 RPM 
         [0069]    Vehicle speed 2.4 mph 
         [0070]    Vehicle tractive force (Torque/R) 1054 lbs. 
         [0071]    Force to weight ratio 0.73 
         [0072]    Force to weight ratio (operator only) 1.24 
         [0073]    Maximum Speed Operation 
         [0074]    Engine RPM 3750 
         [0075]    Engine torque 10.6 ft-lbs. 
         [0076]    Primary clutch pitch diameter 6.36 in. 
         [0077]    Secondary clutch pitch diameter 4.88 in. 
         [0078]    CVT ratio 0.77 
         [0079]    Transaxle (T/A) ratio 18.35 
         [0080]    Total ratio CVT ratio×T/A ratio 14.13 
         [0081]    Input shaft torque 8.2 ft-lbs. 
         [0082]    Axle torque 150 ft-lbs. 
         [0083]    Axle speed 265 RPM 
         [0084]    Vehicle speed 14.0 mph 
         [0085]    Vehicle tractive force (Torque/R) 200 lbs. 
         [0086]    Force to weight ratio 0.14 Force to weight ratio (operator only) 0.24 
         [0087]    The exemplary embodiment of the invention provides a continuously variable transmission having a CVT turn ratio of about 0.77 to 3.1 and a transaxle with a transaxle turn ratio of about 18.35. These ratios, when multiplied, result in a maximum total gear ratio of about 58 and a minimum total gear ratio of about 14. The difference between the maximum total gear ratio and the minimum total gear ratio is about 44. The transaxle turn ratio is about 6 times the maximum CVT turn ratio. The transaxle turn ratio is about 24 times the minimum CVT turn ratio. 
         [0088]    Another exemplary embodiment would have a CVT ratio of between 0.59 and 3.1 and a transaxle ratio of about 24 for a total gear ratio of between about 14 and 74. The difference between the maximum total gear ratio and the minimum total gear ratio is about 60. The transaxle turn ratio is about 7.7 times the maximum CVT turn ratio. The transaxle turn ratio is about 41 times the minimum CVT turn ratio. The transaxle is commercially available from Dana Corporation, Off Highway Systems (Components) Group, Maumee, Ohio, U.S. 
         [0089]    From the foregoing, it will be observed that numerous variations and modifications may be effected without departing from the spirit and scope of the invention. It is to be understood that no limitation with respect to the specific apparatus illustrated herein is intended or should be inferred.