Patent Document

CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application is a Continuation of U.S. patent application Ser. No. 11/040,278, filed on Jan. 20, 2005, issued as U.S. Pat. No. 7,487,954 on Feb. 10, 2009, which claims the benefit of U.S. Provisional Patent Application No. 60/540,105, filed on Jan. 28, 2004, and entitled “Load Control Power Transmission”, the disclosure of each of which are hereby incorporated herein by reference in their entireties. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to winches. More specifically, the present invention relates to transmissions used with winches that are subject to dynamic loading conditions, such as those conditions that arise in a marine environment. 
     BACKGROUND OF THE INVENTION 
     Towing/anchor-handling marine vessels are equipped with winches. When paying out or hauling in wire rope or holding a load stationary, the winches and their wire rope are often subjected to load surges and peaks because of wave action encountered by the vessel. These load surges and peaks can cause the wire rope to fail. 
     The length of wire rope to be paid out from a winch can be significant. Thus, payout of wire rope at normal winch operating speeds can require substantial amounts of time. 
     There is a need in the art for an apparatus and method adapted to minimize the effect of load surges and peaks on winches during payout and haul-in operations in a marine environment. Also, there is need in the art for the ability to perform high speed/horsepower dynamic payout of wire rope in a controlled manner. 
     BRIEF SUMMARY OF THE INVENTION 
     The present invention, in one embodiment, is a transmission used with a winch drum. The transmission includes a fluid cooled clutch coaxially mounted on a drive shaft adapted to drive the winch drum. 
     The present invention, in another embodiment, is a transmission used with a winch drum. The transmission includes a drive shaft, an output shaft, a hydraulic or pneumatic system, a cooling system, a gear coaxially mounted on the output shaft, and an electric motor for powering the gear. The drive shaft is adapted to drive the winch drum and includes a clutch disc extending generally radially outwards from the drive shaft. The clutch disc has a face. The output shaft coaxially surrounds at least a portion of the drive shaft and includes a friction surface extending generally radially inward. The friction surface has a face opposing the face of the clutch disc. The hydraulic or pneumatic system is adapted to bring the faces into contact, and the cooling system is adapted to remove heat from the friction surface via a fluid coolant. 
     The present invention, in another embodiment, is a transmission used with a winch drum. The transmission comprises a drive shaft, an output shaft, an actuation system, and a cooling system. The drive shaft is adapted to drive the winch drum and is operably coupled to a first clutch surface. The output shaft is adapted to be driven by a motor and is operably coupled to a second clutch surface opposing the first clutch surface. The actuation system is adapted to bring the first and second surfaces into contact. The cooling system is adapted to remove heat from at least one of the surfaces via a fluid coolant. 
     The present invention, in another embodiment, is a method of controlling a winch drum transmission equipped with a drive shaft and an output shaft that coaxially surrounds at least a portion of the drive shaft. The drive shaft is adapted to drive a winch drum, and the output shaft is adapted to transfer power from an electric motor to the drive shaft via a hydraulic or pneumatic clutch. The method includes setting a winch load limit, hydraulically or pneumatically causing the clutch to prevent relative displacement between the drive and output shafts when an actual winch load does not exceed the set winch load limit, allowing relative displacement between the shafts when the actual winch load exceeds the set winch load limit, and circulating a fluid coolant through the clutch to remove heat resulting from the relative displacement between the shafts. 
     The present invention, in another embodiment, is a method of performing dynamic payout of wire rope from a winch drum coupled to a transmission. The transmission is equipped with a drive shaft and an output shaft. The drive shaft is adapted to drive the winch drum, and the output shaft coaxially surrounds at least a portion of the drive shaft and is adapted to transfer power from an electric motor to the drive shaft via a hydraulic or pneumatic clutch. The electric motor is electrically connected to an electrical load, such as resistor bank, and the clutch is fluidly connected to a cooling system. Dynamic payout of the wire rope generates energy that needs to be dissipated. In one embodiment, the method includes setting a transition point based on a percentage of the electrical load capacity. In another embodiment, the method includes setting a transition point based on a predetermined electric motor speed. For example, in one embodiment, the predetermined electric motor speed may be based on a percentage of the maximum electric motor speed. The method further includes hydraulically or pneumatically causing the clutch to prevent relative displacement between the shafts when the transition point has not been exceeded, thereby causing all of the energy, generally speaking, to be dissipated through the electrical load, and hydraulically or pneumatically actuating the clutch to allow relative displacement between the shafts when the transition point has been exceeded, thereby causing at least a portion of the energy to be dissipated through the cooling system and the remainder of the energy to be dissipated through the electrical load. 
     The present invention, in another embodiment, is a method of dissipating energy generated by dynamic payout of wire rope from a winch drum. The method includes setting a transition point wherein the responsibility for dissipating the energy transitions from being, generally speaking, the responsibility of a primary energy dissipation system to being shared between the primary system and a supplemental energy dissipation system, dissipating the energy through the primary system when the transition point has not been exceeded, and dissipating the energy through the primary and supplemental systems when the transition point has been exceeded. In one embodiment, the primary system is an electric motor electrically coupled to an electrical load, and the supplemental system is a fluid cooled clutch fluidly coupled to a cooling system. In another embodiment, the primary system is a hydraulic motor fluidly coupled to a hydraulic system, and the supplemental system is a fluid cooled clutch fluidly coupled to a cooling system. 
     While multiple embodiments are disclosed, still other embodiments of the present invention will become apparent to those skilled in the art from the following detailed description, which shows and describes illustrative embodiments of the invention. As will be realized, the invention is capable of modifications in various obvious aspects, all without departing from the spirit and scope of the present invention. Accordingly, the drawings and detailed description are to be regarded as illustrative in nature and not restrictive. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1A  is a starboard elevation of a marine vessel equipped with an anchor-handling/towing winch system. 
         FIG. 1B  is a plan view of the marine vessel illustrated in  FIG. 1A . 
         FIG. 2  is an isometric view of the anchor-handling/towing winch system as viewed from an elevated, port/stem position. 
         FIG. 3  is an isometric view of a load control power transmission as viewed from an elevated, port/stem position. 
         FIG. 3A  is a schematic plan view of an alternative embodiment of the winching system. 
         FIG. 4A  is a sectional elevation along section line AA of  FIG. 3  and through the port clutch, port gear reducer, and outer end of the port drive shaft. 
         FIG. 4B  is a sectional elevation similar to  FIG. 4A , except of an alternative embodiment. 
         FIG. 4C  is a sectional elevation similar to  FIG. 4A , except of an alternative embodiment. 
         FIG. 4D  is a sectional elevation along section line BB of  FIG. 3A  and through a clutch and outer end of a first shaft. 
         FIG. 5  is a flow diagram illustrating the function of the load control power transmission. 
         FIG. 6  is a flow diagram illustrating a dynamic payout process employing the load control power transmission. 
     
    
    
     DETAILED DESCRIPTION 
       FIGS. 1A and 1B  are, respectively, a starboard elevation and a plan view of a marine vessel  1  equipped with the anchor-handling/towing winch system  2  of the subject invention. As illustrated in  FIGS. 1A and 1B , in one embodiment, the winch system  2  is mounted on the deck  3  of the marine vessel  1  with the winch system&#39;s wire ropes  4  feeding towards the stem  5  of the vessel from the winch system  2 . In other embodiments, the winch system  2  is mounted on the deck  3  of a marine vessel  1  so the wire ropes  4  feed towards other parts of the vessel  1 , such as the bow  6 . 
       FIG. 2  is an isometric view of the anchor-handling/towing winch system  2  as viewed from an elevated, port/stern position. As shown in  FIG. 2 , in one embodiment, the winch system  2  includes a port tow drum  10 , a starboard tow drum  11 , an anchor-handling drum  15 , and a load control power transmission  20 . The drums  10 ,  11 ,  15  carry wire rope  4 . 
     The load control power transmission  20  drives and/or brakes the drums  10 ,  11 ,  15  during the winch system&#39;s various in-hauling and payout operations. As shown in  FIG. 2  and explained in the following discussion of  FIGS. 3 and 4 , in one embodiment, the load control power transmission  20  employs a load limiting clutch  65   a ,  65   b  directly on each drive shaft  70   a ,  70   b  to eliminate the effects of motor and power train inertia. Because of each clutch&#39;s location, the speed of its associated motor  45   a ,  45   b , which is operably coupled to a shaft  70   a ,  70   b  and associated drum or drums  10 ,  11 ,  15 , does not have to remain directly proportional to the drum speed during payout. Thus, the load control power transmission  20  allows fall control of the wire rope  4  for normal in-hauling and payout operations, while allowing rapid payout of wire rope  4  during surge or peak load situations, thereby reducing the risk of broken ropes. 
     In one embodiment, the clutches  65   a ,  65   b  are disk or axial type clutches. In one embodiment, the clutches  65   a ,  65   b  are rim type clutches with internal expanding shoes or external contracting shoes. 
     For a more detailed discussion of the load control power transmission  20 , reference is now made to  FIG. 3 , which is an isometric view of the transmission  20  illustrated in  FIG. 2 , as viewed from the same elevated, port/stern position. As shown in  FIG. 3 , in one embodiment, the transmission  20  includes a starboard power assembly  25 , a starboard drive shaft assembly  30 , a port power assembly  35 , and a port drive shaft assembly  40 . The starboard power assembly  25  is operably coupled to the starboard drive shaft assembly  30 . Similarly, the port power assembly  35  is operably coupled to the port drive shaft assembly  40 . 
     As shown in  FIG. 3 , in one embodiment, the power assemblies  25 ,  35  each include an electric motor  45   a ,  45   b , a power shaft  50   a ,  50   b , a brake  55   a ,  55   b , a primary gear reducer  60   a ,  60   b , and a fluid cooled multi-disc clutch  65   a ,  65   b . Each electric motor  45   a ,  45   b  drives a power shaft  50   a ,  50   b  that runs a primary gear reducer  60   a ,  60   b  coupled to a fluid cooled clutch  65   a ,  65   b . Each fluid cooled clutch  65   a ,  65   b , when engaged, transfers the power of its respective electric motor  45   a ,  45   b  to its respective drive shaft assembly  30 ,  40 . As will be explained more fully later in this specification in the discussion of  FIG. 4A , the less a clutch  65   a ,  65   b  is engaged, the greater the amount of slip between its power assembly  25 ,  35  and the respective drive shaft assembly  30 ,  40 . 
     As stated above, one embodiment of the invention employs electric motors  45   a ,  45   b  to drive the winch drums  10 ,  11 ,  15 . However, in other embodiments of the invention, the motors  45   a ,  45   b  are hydraulic motors or internal combustion engines. 
     As illustrated in  FIG. 3 , the drive shaft assemblies  30 ,  40  each include a drive shaft  70   a ,  70   b  supported by drive shaft support bearings  75 . The port drive shaft  70   a  has a port tow drum drive pinion  80   a  and the starboard drive shaft has a starboard tow drum drive pinion  80   b . In one embodiment, as shown in  FIG. 3 , the starboard drive shaft  70   b  has an anchor-handling drum drive pinion  80   c . In another embodiment, the anchor-handling drum drive pinion  80   c  is located on the port drive shaft  70   a . As shown in  FIG. 3 , each pinion  80   a ,  80   b ,  80   c  is paired with a jaw clutch  85   a ,  85   b ,  85   c.    
     As can be understood from  FIGS. 2 and 3 , the port tow drum drive pinion  80   a  interfaces with, and drives, a drive gear on the port tow drum  10 . When the port tow drum  10  is to be driven, the jaw clutch  85   a  engages the pinion  80   a , causing the pinion  80   a  to rotate with the port drive shaft  70   a , thereby driving the port tow drum  10 . When the clutch  85   a  is disengaged from the pinion  80   a , the port tow drum  10  is not driven because the port drive shaft  70   a  is free to rotate within the pinion  80   a.    
     As can also be understood from  FIGS. 2 and 3 , the starboard tow drum drive pinion  80   b  interfaces with, and drives, a drive gear on the starboard tow drum  11 . When the starboard tow drum  11  is to be driven, the jaw clutch  85   b  engages the pinion  80   b , causing the pinion  80   b  to rotate with the starboard drive shaft  70   b , thereby driving the starboard tow drum  11 . When the clutch  85   b  is disengaged from the pinion  80   b , the starboard tow drum  11  is not driven because the starboard drive shaft  70   b  is free to rotate within the pinion  80   b.    
     As can further be understood from  FIGS. 2 and 3 , the anchor-handling drum drive pinion  80   c  interfaces with, and drives, a drive gear on the anchor-handling drum  15 . When the anchor-handling drum  15  is to be driven, the jaw clutch  85   c  engages the pinion  80   c , causing the pinion  80   c  to rotate with the starboard drive shaft  70   b , thereby driving the anchor-handling drum  15 . When the clutch  85   c  is disengaged from the pinion  80   c , the anchor-handling tow drum  15  is not driven because the starboard drive shaft  70   b  is free to rotate within the pinion  80   c.    
     As shown in  FIG. 3 , a center jaw clutch  90  resides between the opposed ends of the drive shafts  70   a ,  70   b . When the center jaw clutch  90  is disengaged, the drive shafts  70   a ,  70   b  are independent of each other and free to rotate at different speeds and different directions, each drive shaft  70   a ,  70   b  being driven by its own power assembly  25 ,  35 . Thus, for example, when the center clutch  90  is disengaged, the port power assembly  35  may drive the port drive shaft  70   a  in one direction to cause the port tow drum  10  to payout its wire rope  4 , while the starboard power assembly  25  may drive the starboard drive shaft  70   b  in the opposite direction to cause the anchor-handling drum or the starboard tow drum to haul-in its corresponding wire rope  4 . 
     As indicated in  FIG. 3 , when the center jaw clutch  90  is engaged, the drive shafts  70   a ,  70   b  essentially become one drive shaft. This allows the power of both power assemblies  25 ,  35  to be applied simultaneously to any one or more of the pinions  80   a ,  80   b ,  80   c  and its corresponding drum  10 ,  11 ,  15 . 
     As indicated in  FIG. 3  and more fully shown in  FIG. 4A , which is a sectional elevation along section line AA of  FIG. 3  and through the port clutch  65   a , port gear reducer  60   a , and outer end of the port drive shaft  70   a , the outer end portion of each drive shaft  70   a ,  70   b  passes through the primary gear reducer  60   a ,  60   b  and terminates within the clutch  65   a ,  65   b . As shown in  FIG. 4A , the primary gear reducer  60   a  includes a housing  100 , a drive gear  105 , a reducer output shaft  110 , support bearings  115  for supporting the reducer output shaft  110  off of the housing  100 , and support bearings  120  for supporting the reducer output shaft  110  off of the drive shaft  70   a.    
     As indicated in  FIG. 4A , the drive shaft  70   a  is supported by the support bearings  75  and is coaxially, rotatably displaceable within the reducer output shaft  110  when the clutch  65   a  is not fully engaged. The reducer output shaft  110  is rotatably displaceable within the housing  100  and supported by the support bearings  115 ,  120 . The drive gear  105  is coaxially mounted on the reducer output shaft  110  and transmits the power from the electric motor  45   a , via the power shaft  50   a , to the reducer output shaft  110 . As will be explained in greater detail later in this specification, the power is then transmitted from the reducer input shaft  110  to the drive shaft  70   a  to a greater or lesser degree, depending on the degree of clutch engagement. 
     As illustrated in  FIG. 4A , in one embodiment, the clutch  65   a  includes a clutch housing  125 , a swivel assembly  130 , a coolant inlet  135 , a coolant outlet  140 , a main hydraulic or pneumatic control pressure line  145 , coolant lines  150 , and branch hydraulic or pneumatic control pressure lines  190 . In one embodiment, where the each clutch  65   a ,  65   b  is a disk or axial type clutch, each clutch  65   a ,  65   b  will also include pressure plate friction surfaces  155  and clutch discs  160 . In one embodiment, a clutch guard  165  encloses all of the aforementioned components of the clutches  65   a ,  65   b , except the pressure line  145  and the coolant inlet  135  and outlet  140 . The clutch housing  125  is secured to the reducer output shaft  110  and is coaxially, rotatably displaceable about the drive shaft  70   a  when the clutch  65   a  is not fully engaged. The swivel assembly  130  is secured to the clutch housing  125 . 
     As indicated in  FIG. 4A , the clutch housing  125  supports pressure plate friction surfaces  155  that are parallel to each other, extend radially inward from the clutch housing  125 , and are secured to the clutch housing  125 . The clutch discs  160  are mounted on the end portion of the drive shaft  70   a , are parallel to each other, and radially extend outward from the shaft&#39;s outer circumference. Each clutch disc  160  is sandwiched between a pair of pressure plate friction surfaces  155 . When the pressure plate friction surfaces  155  are hydraulically or pneumatically actuated by a hydraulic or pneumatic engagement system  170 , they engage the clutch discs  160 . 
     When the pressure plate friction surfaces  155  are less than fully engaged, the clutch discs  160  may rotatably displace relative to the friction surfaces  155 , if a torque exerted on the drive shaft  70   a  exceeds the frictional force between the friction surfaces  155  and the clutch discs  160 . The drive shaft  70   a  would then rotatably displace relative to the reducer output shaft  110 . 
     Conversely, when the pressure plate friction surfaces  155  are fully engaged such that the torque exerted on the drive shaft  70   a  does not exceed the frictional force between the friction surfaces  155  and the clutch discs  160 , the clutch discs  160  are prevented from rotatably displacing relative to the friction surfaces  155  and, as a result, the drive shaft  70   a  does not rotatably displace relative to the reducer output shaft  110 . Consequently, the drive shaft  70   a  and the reducer output shaft  110  rotate together as one shaft. 
     As shown in  FIG. 4A , the coolant inlet  135  and coolant outlet  140  are connected to the swivel assembly  130  to circulate coolant from the cooling system  175  through the clutch housing  125  via the coolant lines  150 . The coolant absorbs and removes heat generated at the friction surfaces  155 . In one embodiment, the fluid coolant is water. In other embodiments, the coolant will be oil, air or other types of fluids. 
     As illustrated in  FIG. 4A , the hydraulic or pneumatic control pressure line  145  runs from the hydraulic or pneumatic actuation system  170  to a connection point on the swivel assembly  130 , which is secured to the clutch housing  125 . The branch hydraulic or pneumatic lines  190  are in fluid communication with the main hydraulic or pneumatic control pressure line  145  and run from the swivel assembly  130  to the clutch housing  125 . The branch hydraulic or pneumatic lines  190  actuate the friction surfaces  155 . Other actuation systems based on magnetic, mechanical or other actuation methods may also be used. 
     While  FIG. 4A  depicts one embodiment of the invention where the drive shaft  70   a  is coaxially positioned within the reducer output shaft  110 , the friction surfaces  155  extend radially inward, and the clutch discs  160  extend radially outward, those skilled in the art will realize that other configurations of the invention may be developed without departing from the spirit of the invention. 
     For example, as illustrated in  FIG. 4B , which is a sectional elevation similar to  FIG. 4A , except of an alternative embodiment, the port clutch  65   a  and the port gear reducer  60   a  have reversed positions and the drive shaft  70   a  is no longer coaxially within the reducer output shaft  110 . Furthermore, the clutch discs  160  extend radially inward from the drive shaft  70   a  or, that is to say, an extension of the drive shaft  70   a , and the friction surfaces  155  extend radially outward from the reducer output shaft  110 , or in other words from a clutch housing  125  mounted on the output shaft  110 . 
     As shown in  FIG. 4B , the coolant inlet  135 , coolant outlet  140 , and main hydraulic or pneumatic control pressure line  145  connect to a swivel assembly  130  on the end of the output shaft  110 . A branch hydraulic or pneumatic line  190  leads from the swivel assembly  130 , through the output shaft  110 , and to the friction surfaces  155 . Coolant supply and return lines  150  run from the coolant inlet  135  and outlet  140 , through the output shaft  110 , and to the friction surfaces  155 . Like the embodiment illustrated in  FIG. 4A , the gear reducer  60   a  causes the output shaft  110  to rotate, which causes the drive shaft  70   a  to rotate to a greater or lesser degree, depending on the degree of clutch engagement. 
     To illustrate another embodiment of the invention, reference is now made to  FIG. 4C , which is a sectional elevation similar to  FIG. 4A , except of an alternative embodiment, wherein the port clutch  65   a  and the port gear reducer  60   a  have reversed positions and the drive shaft  70   a  is no longer coaxially within the reducer output shaft  110 . As shown in  FIG. 4C , the clutch discs  160  extend radially outward from the drive shaft  70   a , and the friction surfaces  155  extend radially inward from the clutch housing  125 , which is attached to the end of the output shaft  110 . 
     As illustrated in  FIG. 4C , the coolant inlet  135 , coolant outlet  140 , and main hydraulic or pneumatic control pressure line  145  connect to a swivel assembly  130  on the end of the output shaft  110 . A branch hydraulic or pneumatic line  190  leads from the swivel assembly  130 , through the output shaft  110 , and to the friction surfaces  155 . Coolant supply and return lines  150  run from the coolant inlet  135  and outlet  140 , through the output shaft  110 , and to the friction surfaces  155 . Like the embodiment illustrated in  FIG. 4A , the gear reducer  60   a  causes the output shaft  110  to rotate, which causes the drive shaft  70   a  to rotate to a greater or lesser degree, depending on the degree of clutch engagement. 
     To illustrate another embodiment of the winching system  2  of the subject invention, reference is now made to  FIG. 3A , which is a schematic plan view of an alternative embodiment of the winching system  2 . As shown in  FIG. 3A , a power shaft  50  extends between a motor  45  and a gear box  60 . A brake  55  is located along the power shaft  50 . A first shaft  70  extends between the gear box  60  and a clutch  65 . 
     As shown in  FIG. 4D , which is a sectional elevation taken along section line BB of  FIG. 3A  and through the clutch  65  and outer end of the first shaft  70 , in extending into the clutch  65 , the first shaft  70  is coaxially surrounded by a second shaft  110  and a first gear  105  mounted on the second shaft  110 . In one embodiment, a clutch housing  125  radially extends from the second shaft  110 . Pressure plate friction surfaces  155  are mounted on the clutch housing  125  and configured to engage clutch discs  160  that radially extend from the first shaft  70 . 
     As can be understood from  FIG. 3A , the first gear  105  drives a second gear  106 , which is mounted on a third shaft  111 . A fourth gear  113  is coaxially pivotally mounted on the third shaft  111  and in engagement with a drum gear  114  on the winch drum  10 . The fourth gear  113  is brought into engagement with the third shaft  111  via a jaw clutch  85  arrangement as previously described in this Detailed Description. When the fourth gear  113  is engaged with the third shaft  111 , it will drive a drum gear  114  and, as a result, the winch drum  10 . 
     To discuss the function of the load control power transmission  20  and its components, reference is now made to  FIGS. 3 ,  4 A and  5 .  FIG. 5  is a flow diagram illustrating the function of the transmission  20 . In operation, the winch operator sets the winch load limit at an operator&#39;s control panel  180  (block  500 ). In other words, the operator sets the clutch  65   a ,  65   b  such that the clutch discs  160  will not rotatably displace relative to the friction surfaces  155 , unless the torque imposed on the clutch  65   a ,  65   b  by the load in the wire rope  4  exceeds the frictional force between the friction surfaces  155  and the clutch discs  160 . In one embodiment, the winch load limit will be based on a percentage of the structural load limit of the winch or a component of the winch (e.g., the structural load limit of the wire rope). 
     The operator then causes the winch to perform a payout or haul-in operation or causes the winch to hold a load in place. If the actual load in the wire rope  4  does not exceed the set load limit (block  510 ), then there is no relative motion between the clutch discs  160  and the friction surfaces  155  (block  520 ). As a result, there is no relative motion between the drive shaft  70   a ,  70   b  and the reducer output shaft  110 , and these shafts operate as one shaft (block  520 ). 
     If the actual load in the wire rope exceeds the set load limit (block  510 ), then there is relative motion between the clutch discs  160  and the friction surfaces  155 , because the clutch discs  160  slip (block  530 ). Consequently, there is relative motion between the drive shaft  70   a ,  70   b  and the reducer output shaft  110  (block  520 ). This situation may arise, for example, during a payout or haul-in procedure when a large wave causes the vessel  1  to surge upwards, suddenly decreasing the slack in the wire rope and causing the wire rope load to peak. Once the actual load in the wire rope returns below the set load limit (block  510 ) (e.g., the vessel  1  rides down the wave and the slack in the wire rope increases), the friction surfaces  155  relock on the clutch discs  160  and the relative motion between the drive shaft  70   a ,  70   b  and the reducer output shaft  110  ceases (i.e., the these shafts again operate as one shaft) (block  520 ). 
     The load control power transmission  20  facilitates dynamic, high speed/high horsepower wire rope payout by providing two modes for dissipating the energy generated during the dynamic payout process. In the first mode, during a dynamic payout, the load control power transmission  20  generates energy via a motor  45   a ,  45   b  and the energy is dissipated at an energy dissipation system  185  connected to the motor  45   a ,  45   b . For example, in one embodiment, the energy is generated at an electric motor  45   a ,  45   b  and the energy is dissipated at an electrical load, such as a resistor bank  185 , electrically connected to the electrical motor  45   a ,  45   b . In the second mode, during a dynamic payout, the load control power transmission  20  generates energy via both an electric motor  45   a ,  45   b  and a clutch  65   a ,  65   b , and the energy is dissipated via the resistor bank  185  coupled to the motor  45   a ,  45   b  and a cooling system  180  coupled to the clutch  65   a ,  65   b.    
     As explained above, in one embodiment of the first mode, the dynamic payout energy may be dissipated at an electrical load (e.g., resistor bank  185 ) coupled to an electric motor  45   a ,  45   b . However, in another embodiment of the first load, wherein the electrical motor  45   a ,  45   b  and the electrical load  185  are replaced with a hydraulic motor coupled to a hydraulic system, the dynamic payout energy is dissipated via the hydraulic system. In either case, in the second mode, the energy generation/dissipation method of the first mode (i.e., the electric motor/electrical load combination or the hydraulic motor/hydraulic system combination) is combined with the energy generation/dissipation capability of the fluid cooled clutch  65   a ,  65   b  coupled to the cooling system  180 . 
       FIG. 6  is a flow diagram illustrating the dynamic payout process. In operation, the winch operator uses the operator&#39;s control panel  180  to set a transition point wherein the load control power transmission  20  shifts from the first mode to the second mode (block  600 ). In other words, the transition point determines when the energy generation/dissipation responsibilities shifts from being, generally speaking, the responsibility of the primary energy generation/dissipation system (i.e., the electric motor/resistor bank combination) to being shared between the primary energy generation/dissipation system and the supplemental energy generation/dissipation system (i.e., the clutch/cooling system combination). 
     In one embodiment, the transition point may be based on a percentage of the resistor bank capacity. For example, in one embodiment, the setting is 66% of the maximum resistor bank dissipation capacity. 
     In one embodiment, the transition point may be based on a predetermined electric motor speed, winch drum speed, and/or torque perceived by the motor. For example, in one embodiment, the predetermined electric motor speed and/or torque may be based on a percentage of the maximum payout motor speed and/or torque. 
     Once the transition point has been set (block  600 ), the operator causes the winch to perform a dynamic payout operation. If the power generated by the electric motor  45   a ,  45   b  does not exceed the setting (e.g., 66% of the maximum resistor bank dissipation capacity or a predetermined payout motor speed) (block  610 ), then the electric motor  45   a ,  45   b  continues to handle the dynamic payout forces by itself (i.e., the electric motor/resistor bank combination is, generally speaking, responsible for the generation and dissipation of all the dynamic payout energy) and there is no relative motion between the clutch discs  160  and the friction surfaces  155  (block  620 ). As a result, there is no relative motion between the drive shaft  70   a ,  70   b  and the reducer output shaft  110 , and these shafts operate as one shaft (block  620 ). Thus, when the load control power transmission  20  is operating in the first mode during a dynamic payout, the speed of the winch drum is controlled by the braking effect of the motor  45   a ,  45   b  and associated electrical load (e.g., resistor bank  185 ). 
     If the power regenerated by the electric motor  45   a ,  45   b  exceeds the setting (e.g., 66% of the maximum resistor bank regeneration dissipation capacity or a predetermined payout motor speed and/or torque) (block  610 ), then the load control power transmission  20  transitions to the second mode and the excess percentage of the resistor bank capacity or the motor speed and/or torque is accommodated by the fluid cooled clutch  65   a ,  65   b  (block  630 ). Specifically, the clutch discs  160  begin to slip allowing relative motion between the clutch discs  160  and the friction surfaces  155  (block  630 ). As a result, there is relative motion between the drive shaft  70   a ,  70   b  and the reducer output shaft  110 , which, in one embodiment, allows the motor  45   a ,  45   b  to slow and decreases the power being sent to the resistor bank  185  (block  630 ). In another embodiment, relative motion between the drive shaft  70   a ,  70   b  and the output shaft  110  at least prevents the motor speed and/or the power being sent to the resistor bank from increasing further. 
     The heat generated by the slipping clutch discs  160  is carried away by the cooling system  175  (block  630 ). Thus, when the load control power transmission  20  is operating in the second mode during a dynamic payout, the speed of the winch drum is controlled by the braking effects of the motor  45   a ,  45   b  and associated electrical load (e.g., resistor bank  185 ) and the slipping discs  160  of the fluid cooled clutch  65   a ,  65   b . Also, in the second mode, the relative motion between the shafts  70 ,  110  allows the speed of the payout to be maintained, although the electric motor  45   a ,  45   b  has been allowed to slow or at least the motor&#39;s speed and/or torque has not continued to increase. 
     Once the power to be dissipated during the dynamic payout process decreases to a level that does not exceed the setting (block  610 ), the friction surfaces  155  fully engage the clutch discs  160  to stop the relative motion between these aspects of the clutch  65   a ,  65   b  (block  620 ). At the same time, the electric motor  45   a ,  45   b , if necessary, speeds up to match the payout speed, and the resistor bank  185  again, generally speaking, becomes responsible for dissipating all of the power being generated by the dynamic payout (block  620 ). 
     In one embodiment, the dynamic payout power being generated by the electric motor  45   a ,  45   b  and sent to the resistor bank  185  is monitored via power sensor means as are known in the art. As the power increases, additional resistors are brought on line (i.e., the electrical load is increased incrementally). Once, the transition point (i.e., a percentage of the electrical load capacity) has been reached, the clutch  65   a ,  65   b  is progressively released and relative rotational displacement between the drive shaft  70   a ,  70   b  and the output shaft  110  progressively increases. As the dynamic payout process continues, the power being sent to the electrical load  185  is continuously monitored and the clutch will be adjusted accordingly. 
     In one embodiment while the system is operating in the second mode, if the power to the electrical load begins to decrease, the power sensors will determine this as an indication that the overall dynamic payout power is decreasing. Consequently, the clutch  65   a ,  65   b  will be actuated to progressively decrease the rotational displacement between the drive and output shafts. If the monitoring system determines that the overall dynamic payout power has decreased to a point that does not exceed the transition point, then the system will begin to transition to the first mode by progressively actuating the clutch to progressively increase the torque perceived by the electrical motor until the system is fully operating in the first mode. 
     As explained above, in one embodiment, as the energy generated during the dynamic payout process causes the set percentage of maximum motor speed or electrical load capacity to be exceeded, the clutch  65   a ,  65   b  begins to slip and the cooling system  175  begins to assume responsibility for at least a portion of the necessary energy dissipation. In other words, the energy dissipation responsibilities transitions from being, generally speaking, the responsibility of the electrical motor  45   a ,  45   b  and its associated electrical load  185 , to being at least partially shared with the clutch  65   a ,  65   b  and the cooling system  175 . 
     However, the responsibilities and sequencing may be reversed. For example, the energy dissipation responsibilities could initially be, generally speaking, the responsibility of the clutch  65   a ,  65   b  and the cooling system  175 . When a set point associated with the clutch (e.g., a percentage of the maximum clutch speed or a percentage of the maximum cooling capacity of the cooling system) is exceeded, the electrical motor  45   a ,  45   b  and its associated electrical load  185  begin to assume at least partial responsibility for energy dissipation. 
     In the event of an emergency stop or drum over-speed condition, the fluid cooled clutch  65   a ,  65   b  is fully applied, along with the drum brakes and the electric motor brakes  55   a ,  55   b , in a controlled sequence. This provides maximum stopping power to the winch. 
     Although the present invention has been described with reference to preferred embodiments, persons skilled in the art will recognize that changes may be made in form and detail without departing from the spirit and scope of the invention.

Technology Category: 7