U.S. patent application number 15/144230 was filed with the patent office on 2016-08-25 for variable ratio gearmotor with interactive ratio control.
The applicant listed for this patent is Richard C. Raney. Invention is credited to Richard C. Raney.
Application Number | 20160245383 15/144230 |
Document ID | / |
Family ID | 39136521 |
Filed Date | 2016-08-25 |
United States Patent
Application |
20160245383 |
Kind Code |
A1 |
Raney; Richard C. |
August 25, 2016 |
VARIABLE RATIO GEARMOTOR WITH INTERACTIVE RATIO CONTROL
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.
Inventors: |
Raney; Richard C.; (Round
Rock, TX) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Raney; Richard C. |
Round Rock |
TX |
US |
|
|
Family ID: |
39136521 |
Appl. No.: |
15/144230 |
Filed: |
May 2, 2016 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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11510718 |
Aug 28, 2006 |
9328810 |
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15144230 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F16H 15/20 20130101;
F16H 37/0853 20130101; F16H 37/086 20130101; F16H 15/18
20130101 |
International
Class: |
F16H 37/08 20060101
F16H037/08; F16H 15/18 20060101 F16H015/18 |
Claims
1-19. (canceled)
20. A transmission comprising a drive shaft, said drive shaft
having an inner shaft and an outer shaft, a friction disk
nonrotatably mounted to said outer shaft, a shifter connected to
said friction disk, said shifter comprising an actuating system to
cause movement of the friction disk along the outer shaft, at least
two cones engaging the periphery of said friction disk, and a
differential gear set combining the rotation of said inner and
outer shaft.
21. The transmission of claim 1, further comprising an output shaft
connected to said differential gear set.
22. The transmission of claim 1, including a clutch for connecting
said inner shaft to said output shaft.
23. An infinitely variable ratio gearbox comprising a drive motor
and a control system for operation of a transmission having a
coaxial shaft having an inner and outer rotatable shaft connected
to a differential gear set to combine the rotation of the inner and
outer shaft to power an output shaft connected to a load, wherein
the transmission is arranged to provide infinite ratio at zero
output speed whereby the drive motor can be started without
load.
24. The infinitely variable ratio gearbox of claim 4, wherein the
control system is configured to adjust the gearbox ratio in a
continuous manner in response to a motor load ampere signal
supplied to the control system so as to provide a smooth,
controlled start up and controlled operation of the load.
25. The infinitely variable ratio gearbox of claim 4, wherein the
control system maintains the drive motor at substantially its most
efficient speed and load while continuously adjusting the gearbox
ratio for optimizing output torque and speed for the load for a
given application.
26. The infinitely variable ratio gearbox of claim 5, wherein the
control system controls the output speed of the drive motor either
by direct shift of the gearbox ratio or by adjusting the motor
ampere load reference signal.
27. The infinitely variable ratio gearbox of claim 5, wherein the
control system controls the output speed of the drive motor by
adjusting the motor ampere load reference signal, wherein 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.
28. The infinitely variable ratio gearbox of claim 5, wherein the
control system is configured to cause automatic adjustment of the
ampere reference signal.
29. The infinitely variable ratio gearbox of claim 4, wherein the
control system is configured to cause automatic start up or shut
down of the system.
30. The infinitely variable ratio gearbox of claim 4, wherein the
drive motor is a poly-phase alternating current motor.
31. The infinitely variable ratio gearbox of claim 4, wherein the
transmission comprises a drive shaft, said drive shaft having an
inner shaft and an outer shaft, a friction disk nonrotatably
mounted to said outer shaft, a shifter connected to said friction
disk, the shifter comprising an actuating system to cause movement
of the friction disk along the outer shaft, and at least two cones
engaging the periphery of the friction disk, wherein the control
system is configured to continually position the friction disk via
the actuating system to substantially maintain a set motor
load.
32. The infinitely variable ratio gearbox of claim 4, wherein the
load is a pumping unit.
33. A method of controlling drive to a load comprising providing an
infinitely variable ratio gearbox comprising a drive motor and a
transmission having a coaxial shaft having an inner and outer
rotatable shaft connected to a differential gear set to combine the
rotation of the inner and outer shaft to power an output shaft
connected to a load, controlling the transmission to provide
infinite ratio at zero output speed whereby the drive motor can be
started without load.
34. The method of claim 14, further comprising controlling the
gearbox ratio in a continuous manner in response to a motor load
ampere conditions so as to provide a smooth, controlled start up
and controlled operation of the load.
35. The method of claim 14, further comprising controlling the
drive motor to operate at substantially its most efficient speed
and load while continuously adjusting the gearbox ratio for
optimizing output torque and speed for the load for a given
application.
36. The method of claim 14, further comprising controlling the
output speed of the drive motor either by direct shift of the
gearbox ratio or by adjusting a motor ampere load reference
signal.
37. The method of claim 14, further comprising controlling the
output speed of the drive motor by adjusting the motor ampere load
reference signal, wherein 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.
38. The method of claim 14, further comprising controlling the
operation of the gearbox to cause automatic start up or shut down
of the gearbox.
39. The method of claim 14, further comprising providing the
transmission with a drive shaft having an inner shaft and an outer
shaft, a friction disk nonrotatably mounted to the outer shaft, a
shifter connected to said friction disk, the shifter comprising an
actuating system to cause movement of the friction disk along the
outer shaft, and at least two cones engaging the periphery of the
friction disk, and controlling the position the friction disk via
the actuating system to substantially maintain a set motor load.
Description
BACKGROUND OF THE INVENTION
[0001] 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.
[0002] 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.
[0003] 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.
[0004] 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.
[0005] 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
[0006] 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.
[0007] 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.
[0008] 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.
[0009] 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.
[0010] 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
[0011] FIG. 1 is a partially sectioned side view of an apparatus
formed according to the present invention;
[0012] FIG. 2 is a hydraulic schematic illustrating the hydraulic
circuit of an embodiment of the invention;
[0013] FIG. 3 is an electrical schematic illustrating the power
circuits and control circuits of an embodiment of the
invention;
[0014] FIG. 4 is a schematic illustration of the control module for
an embodiment of the invention;
[0015] FIG. 5 is a partially sectioned side view of an alternate
embodiment of the invention; and
[0016] FIG. 6 is a cross sectional view of an alternative
embodiment of the cone.
DETAILED DESCRIPTION OF THE INVENTION
[0017] 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.
[0018] 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.
[0019] A shifting collar 6 connects to hub 5 by means of a bearing
such as a thrust bearing within collar 6. Collar 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.
[0020] 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.
[0021] 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.times.(Rated Load/Applied Load).sup.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.degree.) 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.
[0022] 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.
[0023] 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.
[0024] 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.
[0025] A set of differential gears combines the torque of the
shafts 2 & 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.
[0026] 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.
[0027] 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.
[0028] 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.
[0029] 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.
[0030] 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.
[0031] 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 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.
[0032] Pump motor 55 is controlled by a 3-phase starter/contactor
which comprise the usual coil 126, power contacts 126p and control
contacts 126a1, 126a2 and 126a3 as shown.
[0033] 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 123p and control contacts 123a1, 123a2
and 123a3; a Y contactor having the usual coil 120, associated
power contacts 120p and a normally closed control contact 120a; and
a DELTA contactor having the usual coil 121, associated power
contacts 121p and control contacts 121a and 121c. The pump
starter/contactor 126 and the main contactor 123 include safety
overload tripping contacts 78 and 79, respectively.
[0034] 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.
[0035] A timing relay 124, having associated delay contacts 124od
and 124cd, 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 123a2 as shown. Also, at this time Y
contactor 120 will be energized through the normally closed
contacts 124od and 121a. The power contacts 120p 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 124cd and 120a, as shown. The
power contacts 121p 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.
[0036] The starter coil 123 is accompanied by a timing relay 125
equipped with both a delay contact 125d and an instant contact
125i. In a starting sequence, relay 125 will be energized first and
its instant contact 125i 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 125d 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.
[0037] The control circuit may also be equipped with a pilot relay
127 having an auxiliary contact 127a 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.
[0038] 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.
[0039] 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.
[0040] 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.
[0041] FIG. 4 schematically illustrates a control module 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.
[0042] 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.
[0043] 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.
[0044] Module 70 receives seven input signals and produces one
output signal plus output servo power. For instance, voltage
through normally closed contact 121c 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.
[0045] Of course, when motor 53 is running, the normally closed
contact 121c 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 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.
[0046] 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.
[0047] 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.
[0048] 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
126a3 opens to de-energize valve 86 when contactor 126 is
de-energized. Thus, valve 86 closes to hold clamping pressure on
pistons 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, 125i 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 121c 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.
[0049] 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.
[0050] 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.
[0051] 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.
[0052] 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.
[0053] 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.
[0054] 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.
[0055] 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.
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