Abstract:
A variable ratio gearbox is combined with a drive motor and an interactive control system. A coaxial shaft connects to a differential gear set to power an output shaft. The transmission has a drive shaft with an inner shaft and an outer shaft. A friction disk nonrotatably mounts to said outer shaft and a shifter connected to the friction disk moves the friction disk along the outer shaft. At least two cones engage the periphery of the friction disk and a differential gear set combines the rotation of the inner and outer shaft.

Description:
BACKGROUND OF THE INVENTION 
     The disclosure pertains to geared drives commonly referred to as “gear motors”. Gear motors drive a wide range of different industrial machines, such as pumps, conveyors, rock crushers, rotary kilns, hoists, and some types of vehicles. 
     Many of these types of machines, such as oil well pumping units, air conditioning compressors, loaded ore conveyors, sand and gravel conveyors, loaded rotary kilns and loaded hoists require relatively high driving torque for starting and bringing up to operating speed. Typically, such machines are started and driven by electric motors acting through fixed ratio gearing and are therefore subjected to high starting loads. Hence, to handle the starting loads, such driving motors are usually sized larger than is required for steady running. It is well known that electric motors are most efficient when sized to be near full load during steady operation. Thus, the high starting loads combined with less than optimum running efficiency causes considerable waste of electric power. 
     Also, some of these types of machines, such as oil well pumping units, rock crushers, hoists and the like, experience widely varying torque loads during normal operation. Typically, electric motors used for driving such varying loads are the NEMA D “high-slip” type in order to withstand the load variation without overheating or having to be extremely oversized. Of course, both “high-slip” motors and oversized motors are considerably less efficient than correctly sized “low-slip” or “premium efficiency” motors. 
     In addition, some of these types of machines, such as oil well pumping units, gas compressors, boiler feed pumps and hoists must operate at variable output speeds to accommodate changing operating requirements. Typically, such speed changes are accomplished by electronic speed variation of the drive motor or by hydraulically varying the speed relation between the drive motor and the load. Both of these methods produce undesirable inefficiencies. 
     Therefore, it is an object of the invention to provide a gear motor having a suitably variable mechanical ratio for driving a wide variety of different machines at suitable speeds and/or variations of speeds in a highly efficient manner; to reduce and/or eliminate inherent motor starting loads; and to facilitate the use of premium efficiency motors instead of other, less efficient types. 
     SUMMARY OF THE INVENTION 
     The invention is a combination of an infinitely variable ratio gearbox with an drive motor and an interactive control system. A coaxial shaft connects to a differential gear set to power an output shaft. 
     The gearbox is arranged to provide infinite ratio at zero output speed whereby the drive motor can be started without load. The control system adjusts the gearbox ratio in a continuous or “step-less” manner in response to a motor load ampere signal so as to provide a smooth, controlled start up and controlled operation of a load. The control system maintains the drive motor at its most efficient speed and load while continuously adjusting the gearbox ratio for optimum output torque and speed for the particular load application. 
     Output speed can be adjusted either by direct shift of the gearbox ratio or by adjusting the motor ampere load reference signal. For instance, a higher ampere setting will produce higher output acceleration and/or speed while a lower ampere setting will produce lower output acceleration and/or speed. The control system may be arranged for automatic adjustment of the ampere reference signal and/or automatic start up or shut down of the system. 
     When used in electric powered vehicles, the invention can enable higher output torque at low speeds and higher maximum output speeds for the given drive motor size than conventional methods. 
     For industrial applications, the drive motor is preferably a poly-phase alternating current motor of the highest efficiency design and is maintained at its most efficient speed. Other types of electric motors may be used for particular applications. For example, direct current motors may be preferred for use in vehicles, especially for battery powered vehicles. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a partially sectioned side view of an apparatus formed according to the present invention; 
         FIG. 2  is a hydraulic schematic illustrating the hydraulic circuit of an embodiment of the invention; 
         FIG. 3  is an electrical schematic illustrating the power circuits and control circuits of an embodiment of the invention; 
         FIG. 4  is a schematic illustration of the control module for an embodiment of the invention; 
         FIG. 5  is a partially sectioned side view of an alternate embodiment of the invention; and 
         FIG. 6  is a cross sectional view of an alternative embodiment of the cone. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     In  FIG. 1 , an embodiment of the invention is shown in a typical environment wherein the invention is used to drive a machine such as an oil well pumping unit. 
     The transmission has a central tractionally driven member or disk  1  mounted upon and with a hollow supporting coaxial shaft  2 . All elements of the apparatus are located within a housing  3  and cap  15 . The shaft  2  is supported on appropriate bearings, such as bearing set  4  at the front end of shaft  2  and an appropriate bearing at its opposite end mounted within the hub of a planet gear carrier  28 . Disk  1  is nonrotatably mounted upon a concentric hub  5  by any conventional means such as a spline conforming to and engaging with a straight spline portion of shaft  2 . Hub  5  is so arranged to allow disk  1  to travel axially along shaft  2  in a low friction manner while simultaneously transmitting torque cooperatively. Although disk  1  is shown as a single disk, it may be of a multiple disk structure. 
     A shifting collar  6  connects to hub  5  by means of a bearing such as a thrust bearing within collar  6 . Collar  6  is arranged to move axially along shaft  2  to thereby control axial position of disk  1  while allowing free rotation of disk  1  and shaft  2  relative thereto. 
     An actuator controls the axial position of hub  5  and disk  1 .  FIG. 1  depicts one type of actuator having a floating lever  116  pivotally mounted to collar  6  and pivotally anchored by a fulcrum link  117  which is in turn pivotally fastened to cap  15  generally as shown. The driven end of lever  116  is pivotally attached to actuating rod  51  of a servo actuator  50 . As shown herein, actuator  50  is electrically driven, but could be hydraulically driven. 
     A plurality of conical rotors  7  are symmetrically positioned circumferentially about disk  1  so that the inwardly facing sides of the cones  7  are parallel to shaft  2  and in frictional engagement with the rim of disk  1 . In the present embodiment, five cones  7  may be provided about the periphery of disk  1 , though only two are shown. For tractionally driven elements such as disk  1 , the fatigue life of the driven element as well as the bearings on which they may be supported, can be calculated using the following commonly known formula:
 
Life (hours)= K /RPM×(Rated Load/Applied Load) 3  
 
From this, it should be noted that the applied radial, or normal, traction contact load, which is proportional to torque load, must be held to relatively light limits to prevent rapid fatigue failure. Further, applied load has a much greater effect on fatigue life than does rotational speed of the elements. The torque output from disk  1  is also proportional to the applied traction contact normal load and number of tractional contact points. Again, the optimum number of tractional drive members or cones  7  may be therefore chosen for the particular environment and application of the power transfer system. In the present invention, the preferred embodiment utilizes the maximum number of contact points to provide maximum power output, with final output speed predetermined by a differential gear assembly which will be described hereafter. The disk member  1  as well as the tractional drive member or cone  7  are designed and arranged to sustain extremely high rotating speeds without causing premature failure. The cones  7  of this embodiment are fifteen degree (15°) cones each having approximately four to one (4:1) diametrical ratio, although other embodiments might include cones of a different angle, size and ratio configuration. In this embodiment, the major outside diameter of each cone  7  is approximately equal to the outside diameter of disk  1 . U.S. Pat. No. 6,001,042, incorporated herein by reference, shows the cone assembly in more detail.
 
     Each cone  7  includes a concentric shaft extending from each end supported by bearings  8  and  9 . A drive gear  10  is provided at its front end, with a thrust bearing  12  supporting the shaft at its front end, as shown. Bearings  8  and  9  are, preferably, needle-type roller bearings capable of sustaining extremely high rotating speeds and relatively high radial loads. Bearings  8  and  9  are mounted in bearing blocks  13  and  14 , respectively, which are in turn fastened to a cone assembly cap  15  attachable to the main housing  3 . A concentric piston  16  is arranged to abut thrust bearing  12  rearward and to be sealed by means of an o-ring  17  within a forcing cylinder  18 . The cylinder  18  in turn abuts the end plate  23  of cap  15  forwardly so that piston  16  applies axial thrust through bearing  12  to cone  7  whenever fluid pressure is directed through a port  19  into cylinder  18  and against piston  16 . In this manner, clamping force, or normal force, between cones  7  and disk  1  is achieved to prevent slippage. Bearings  8  and  9 , and the mating shaft journals of cone  7 , are arranged to allow slight but adequate axial movement of cone  7 . Drive gear  10  may be keyed on its shaft and is retained on its shaft by axial force through bearing  12 . 
     As shown in  FIG. 1 , an input ring gear  21  is provided in common mesh with drive gears  10  at a step-up ratio of 1:5 and is supported in bearing set  22  in association with end plate  23 . Bearing set  22  is preferably of the type capable of supporting both radial and axial loads and is retained within plate  23  by a retaining plate  24  and by nut  26  which is in turn installed on the hub of gear  21 . Bearing set  4  comprises a pair of opposed thrust bearings arranged to support and control axial force and movement of shaft  2  and is mounted concentrically within gear  21 . Gear  21  is arranged to be coupled to and driven by a suitable driver, such as electric motor  53 . 
     Drive motor  53  may be a flange mounted type which may be mounted to plate  23  and cap  15  by means of an adapter ring  52 . A sleeve type coupling  25  is keyed to the shaft  54  of drive motor  53  and is splined to the hub of gear  21  and to the end of an internal coaxial shaft  27 . The shaft  27  is provided in common mesh with coupling  25  and extends through coaxial bores in other elements of the mechanism rearward where it meshes by splines with planet carrier  28 . 
     A set of differential gears combines the torque of the shafts  2  &amp;  27 . One example of the gear set comprises output planet gears  43  which spin on bearings mounted on axially disposed spindles secured to carrier  28 . Planet gears  43  are in common mesh with sun gear  45  which is in turn keyed onto shaft  2 . Ring gear  44  is also in common mesh with planet gears  43  and the hub of gear  44  rotates on a bearing mounted on the neck of carrier  28 . The hub of gear  44  is supported in bearing set  46  and is integral to output shaft  47 . Shaft  47  will of course be coupled to drive a load. In this embodiment, the diametrical ratio between ring gear  44  and sun gear  45  is 4.5:1, although other ratios are of course possible. 
     Parallel or “split torque” transmission of torque is accomplished by the transmission of one torque reaction through gear  21 , gears  10 , cones  7 , disk  1 , shaft  2  and sun gear  45  to the differential planet gears  43 ; and then another torque reaction through shaft  27 , through carrier  28  to planet gears  43 , which is the other parallel torque path. The differential action of planet gears  43  combines the two separate torques and directs a single torque to ring gear  44 . In this manner, the main torque load is transmitted through shaft  27 , while the torque load through cones  7  and disk  1  is minimized, yet facilitating the variation of ratio by axial movement of disk  1 . A position transducer  88  is attached to the pivot end of rod  51  to provide a ratio position signal for disk  1 . 
     Although not shown herein, the proposed load, being an oil well pumping unit, may include a spring applied and electrically releasable brake for aiding in safely stopping operation whenever power is disconnected. 
     Also, as shown in  FIG. 5 , any conventional clutch  90  may be installed within gear  44  and between gear  44  and carrier  28  to selectively connect gear  44  to carrier  28  in certain applications of the invention. Clutch  90  may be a commercially available type with known characteristics and engagement methods, so its details are not shown here. Its usage will be further explained later. 
       FIG. 2  schematically illustrates a hydraulic system for supplying controlled pressure to pistons  16 . The pressure system comprises a pump  55  which would, preferably, be driven by a small drive motor separate from main motor  53 . Pump  55  would, preferably, also be a variable displacement type to minimize energy consumption. Pump  55  will intake hydraulic fluid from a sump  60  and supply pressure through line  56  to a pressure control valve  89 . Gauge  57  is also connected to line  56  to read out pressure. Valve  89  provides controlled pressure through line  87  through shut off valve  86  to line  58  which is connected to all ports  19  and thus to pistons  16 . Shut-off valve  86  should be the normally closed type which opens when energized and closes when de-energized. Also, valve  86  should be equipped with an adjustable bleed-off means so that it can bleed-off pressure in a desired manner. A pressure switch  81  is installed in line  58  to indicate when pressure is up for operation and gauge  59  is connected to line  58  to read out pressure. Valve  89  has a usual drain line into sump  60 . 
     In addition, although not shown herein, a lubrication system will be required to both lubricate and cool various moving components, such as bearings, gears, cones and disks. Since such a lubrication system or methods are commonly known by those skilled in the art, it will not be shown herein. 
       FIG. 3  schematically illustrates both the power circuits and control circuits relating to the systems previously described. Transformer  77  converts the high power line voltage to 120 volt AC control power. Main motor  53  and pump motor  55  are shown connected to the power circuit. Also, a bridge rectifier  84  is connected to the power circuit to supply DC power to a brake coil  85 . Usage of brake coil  85  will be explained later. 
     Pump motor  55  is controlled by a 3-phase starter/contactor which comprise the usual coil  126 , power contacts  126   p  and control contacts  126   a   1 ,  126   a   2  and  126   a   3  as shown. 
     Main motor  53  is controlled by a Y/Delta starting system which comprises a main contactor having the usual coil  123  and associated power contacts  123   p  and control contacts  123   a   1 ,  123   a   2  and  123   a   3 ; a Y contactor having the usual coil  120 , associated power contacts  120   p  and a normally closed control contact  120   a ; and a DELTA contactor having the usual coil  121 , associated power contacts  121   p  and control contacts  121   a  and  121   c . The pump starter/contactor  126  and the main contactor  123  include safety overload tripping contacts  78  and  79 , respectively. 
     A neutral position switch  80  is located in the control circuit so that contactor  123  cannot energize unless actuator  50  and disk  1  are at neutral position. Based on the various component ratios presented herein, the neutral position for disk  1  is near the large end of cones  7  which yields a combined ratio of the mechanism equal to infinity and thus an output speed at shaft  47  equal to zero. Conversely, for lowest combined ratio and maximum output speed at shaft  47 , the position of disk  1  will be near the small end of cones  7 . 
     A timing relay  124 , having associated delay contacts  124   od  and  124   cd , provides a delayed transfer from Y motor connection for starting to a DELTA motor connection for running. For instance, when main contactor  123  is energized, timer  124  will be energized through the auxiliary contact  123   a   2  as shown. Also, at this time Y contactor  120  will be energized through the normally closed contacts  124   od  and  121   a . The power contacts  120   p  will connect motor  53  for Y configuration for low ampere starting. Timer  124  will be set to time out after motor  53  has achieved full speed so that Y contactor  120  will be de-energized and D contactor will  121  be energized through auxiliary contacts  124   cd  and  120   a , as shown. The power contacts  121   p  connect motor  53  for DELTA configuration for full power running. Contactors  120  and  121  are mechanically interlocked, as denoted by hidden lines, so they cannot both be energize at the same time. 
     The starter coil  123  is accompanied by a timing relay  125  equipped with both a delay contact  125   d  and an instant contact  125   i . In a starting sequence, relay  125  will be energized first and its instant contact  125   i  will energize brake coil  85  before contactor  123  energizes motor  53 . After coil  85  has sufficient time to begin releasing the brake (approximately 0.5 second), the delay contact  125   d  will energize contactor  123  which then energizes motor  53  and the subsequent Y/DELTA sequence. Actually, Y/DELTA starting systems are well known to those skilled in the art. 
     The control circuit may also be equipped with a pilot relay  127  having an auxiliary contact  127   a  and arranged to be energized through a master ON-OFF switch  128 . The circuit may also include a MANUAL START switch  129  and a MANUAL STOP switch  130 . 
     An auto-start contact  75  and an auto-stop contact  76  may be included for automatic starting and stopping the system. Contacts  75  and  76  will be located remotely in a conventional “pump-off control unit” normally found on oil well pumping units. Also, the pressure switch  81  previously mentioned is included in the circuit to prevent start up of main motor  53  until oil pressure and lubrication is up to a preset amount. In addition, thermostatic switches  82  and  83  are included to prevent running of main motor  53  when the oil temperature is not within preset limits. 
     During any start up sequence, the gearbox ratio is infinity and the torque at shaft  47  is zero until motor  53  reaches operating speed and is switched from Y connection to DELTA connection. Thus, motor  53  always starts at zero load. 
     A set of current pick up coils  92  are installed on the power conductors to motor  53  to provide a motor load signal through line  68  to the control system to be explained later herein. 
       FIG. 4  schematically illustrates a control module  70  and its associated circuits. Module  70  includes a programmable microprocessor having multiple input and output signal capability. Module  70  also includes a servo drive power supply controllable by the microprocessor. 
     The face of module  70  may include an ammeter  71  to provide read out of motor load based on the load signal provided through line  68 . Module  70  is also equipped with a selector switch  72  for selecting between on/off operating mode and slow-down/speed-up mode. For instance, in on/off mode, a pumping unit will be stopped when a pump-off condition is signaled and restarted automatically by a time clock or the like. In slow-down/speed-up mode, the pumping unit will merely slow down a preset amount when a near pump-off condition is detected and speeded up a preset amount after a preset amount of time to keep the pumping rate just below well production rate. 
     Module  70  is also equipped with a manual adjustment  73  for setting the ampere rate which controls the servo driver. In addition, module  70  is equipped with a manual adjustment  74  for setting the rate of ampere rise during startup, which determines the output starting torque and acceleration. Adjustment  74  will be set to provide an optimum acceleration rate which does not overload the mechanism of the gear box or the mechanism of the pumping unit. 
     Module  70  receives seven input signals and produces one output signal plus output servo power. For instance, voltage through normally closed contact  121   c  through line  69  signals module  70  to position servo actuator  50  for zero motor load and to an infinite ratio or “neutral” position. Servo drive power is directed through line  62  while the servo/ratio position signal is received through line  64 . 
     Of course, when motor  53  is running, the normally closed contact  121   c  shown in  FIG. 4  is open so that the servo driver can position actuator  50  according to the ampere load setting. Thus, based on the motor load signal received through line  68 , the servo driver will continually position actuator  50 , and thus the ratio of disk  1 , so as to maintain the set motor load. During this time, based on the combined signals from line  64  for ratio position, from line  65  for oil temperature and from line  68  for motor load, clamp pressure signal through line  63  is modulated to maintain optimum pressure to pistons  16  to prevent slippage between disk  1  and cones  7  at the prevailing torque load. 
     Such applications such as an oil well pumping unit experience drastic load fluctuations during each cycle of the pumping unit during operation. Hence, in working to keep motor load at the set ampere rate, the servo driver of module  70  will continually position actuator  50  and disk  1  in phase with the changing torque load of the pumping unit. For instance, as torque load at shaft  47  increases, disk  1  will be continually shifted to a higher ratio as shaft  47  slows and the torque load on motor  53  holds steady. As torque load at shaft  47  decreases, disk  1  will be continually shifted to a lower ratio as shaft  47  speeds up and the load on motor  53  remains steady. Although the pumping unit will speed up and slow down during each stroke cycle, a given motor ampere setting will produce a given overall stroke rate. The motor ampere setting is increased to increase the stroke rate and the ampere setting is decreased to reduce the stroke rate. 
     When operating in the slow-down/speed-up mode, selector  72  will be in the slow-down position. The pumping unit will normally be equipped with a conventional “pump-off controller” (not shown) which can supply an auto ampere decrease signal through line  66  whenever a near pump-off condition is detected. Module  70  will be pre-programmed to automatically reduce the ampere setting a preset amount whenever an auto ampere decrease signal is received. Thus, the pump stroke rate will be reduced so as to not pump-off the well, which might cause pump cavitation and possible damage. In addition, the conventional “pump-off controller” may be equipped with a timing device to periodically send an auto ampere increase signal to module  70  which may be programmed to slightly increase the ampere rate and thereby the stroke rate. In this manner, the stroke rate may be adjusted to an optimum rate without causing an actual pump-off condition. 
     When operating in the on/off mode, selector  72  will be in the on/off position. In this mode, whenever the conventional “pump-off controller” detects a near pump-off condition it can send an auto stop signal through line  91  to module  70  which then initiates a pump shut down sequence. In a shut down, the auto stop contact  76 , shown in  FIG. 2 , (included in the “pump-off controller”) will open which de-energizes contactor  126  which in turn de-energizes contactors  123  and  121  and relay  125 . Contact  126   a   3  opens to de-energize valve  86  when contactor  126  is de-energized. Thus, valve  86  closes to hold clamping pressure on pistons  16  to prevent slippage between cones  7  and disk  1  during shut down. Valve  86  should be adjusted to allow sufficient bleed off of pressure after shut down so that disk  1  can safely move axially without rotating. Also at this time,  125   i  opens to de-energize brake coil  85  setting the brake on the pumping unit. The brake will be adjusted to bring the pumping to a soft stop and prevent roll-back. Also, during this time, contact  121   c  shown in  FIG. 4  closes to provide a signal through line  69  which directs module  70  and the servo driver therein to position actuator  50  and disk  1  to neutral position. Thus, the pumping unit, motor  53  and the entire mechanism will be brought to a stop. 
     Likewise, in the event of an unplanned shut down, such as loss of electric power during a lightning storm, shut down will occur in the same manner as previously described, except that module  70  and the servo driver will not return disk  1  to neutral position until electric power is restored. 
     To automatically restart, the conventional “pump-off controller” will include the auto-start contact  75 , shown in  FIG. 2 , which may close at a preset time. Start up is thus achieved in the manner previously described. 
     In some applications, other than and unlike oil well pumping units, the load needs to run at a steady speed when up to speed and in operation. In such steady speed applications, an arrangement such as clutch  90  shown in  FIG. 5  previously mentioned, can be utilized to connect carrier  28  to the hub of gear  44  after gear  44  has been brought up to operating speed. In this manner, the driving torque from motor  53  can be directly transmitted through shaft  27 , to carrier  28 , through clutch  90  to ring gear  44 , and out through output shaft  47  without need of any torque load on disk  1  during steady running. 
     In order to utilize clutch  90 , the ratios of gears  21 ,  10 ,  45 ,  43  and  44 , as well as the cone-to-disk ratio, must be arranged so that when disk  1  reaches the small end of cones  7  during acceleration, the speed of ring gear  44  is equal to the speed of carrier  28 , which enables smooth engagement of clutch  90 . 
     When clutch  90  is thus engaged, module  70  and the servo driver may be programmed to keep disk  1  at a synchronized position which prevents torque application to disk  1 . Also, during this time the clamping pressure through line  58  to pistons  16  can be reduced to a minimum amount whereby disk  1  and cones  7  and associated bearings are not subjected to any consequential stress or wear during steady operation of a load. 
     An alternative cone, depicted in  FIG. 6 , has a circumfrentially extending lightening groove  110  formed in the end face of the cone and communicating with lightening holes  111 . While any number of holes may be used, it is envisioned that each cone will have six. The lightening holes  111  meet at a common point and a cooling oil passage  112  extends from this point to the opposite end face of the cone. 
     While the invention has been described with reference to preferred embodiments, a wide range of sizes, variations, alterations or modifications of the invention are possible without departing from the intent and scope of the invention.