Patent Abstract:
An actuator and control algorithm which provide an operator with the ability to intuitively and responsively maneuver heavy work-pieces with ease and precision. The structure of the apparatus may provide a hoist with a compliant sensing system to measure the weight of the payload. The compliant sensing system may result in smaller dead-bands than are realizable with traditional force sensing methods. At the command of the user, the control algorithm may switch between two distinct operational modes: float mode and manual mode. In float mode, the hoist actively counterbalances the weight of the load, allowing it to feel substantially weightless in the operator&#39;s hands. The operator can apply forces directly to the payload to accelerate it in the desired vertical direction. Because of the small dead-band realized with compliant sensing, the payload may be highly responsive to the operators force inputs. As a result, the payload may be intuitively maneuvered at very high speeds, as well as very low speeds. Alternately, the operator may choose to operate in manual mode. While in manual mode, the hoist operates like traditional lifting hoists, responding to velocity commands issued from a remotely controlled pendant.

Full Description:
This application claims priority to the Provisional Patent Application No. 60/333,610 submitted Nov. 27, 2001 by Christopher J. Morse, Benjamin T. Krupp, Jerry E. Pratt and Aaron G. Flores using U.S. Express Mail No. ET402315547US, which is hereby incorporated by reference in its entirety. 

   FIELD OF THE INVENTION 
   The present invention is directed generally to an actuator. Specifically, the present invention is directed to a hoist that is responsive to force inputs. 
   BACKGROUND OF THE INVENTION 
   A traditional hoist consists of a motor connected to a spool, which is used to wind a cable up and down to move a payload vertically. Typically, these devices are remotely controlled with various up/down buttons on a pendant. By attaching the payload to the end of the hoist cable, a human operator can raise and lower heavy payloads by simply pressing the pendant&#39;s buttons. This type of hoist can be found in many manufacturing operations requiring movement of payloads that are too heavy for human operators. Though traditional hoists are indispensable for their tremendous load lifting capabilities, their slow, non-variable speeds, and remote controlled operation can be less than ideal for many manufacturing applications. 
   Consider a simple automotive assembly procedure such as placing an engine block on its engine mounts. During the assembly sequence, the operator raises an engine block from the factory floor, up and over the front fender, into the engine bay and onto the motor mounts. An extremely wide range of speeds would be desirable for this task. While moving from the factory floor to up and over the front fender, it would be desirable to move at quick human-like speeds. Slower speeds would be desired while lowering the engine block into the car&#39;s engine bay. Finally, extremely low speeds with regular changes in direction would be most suitable when precisely placing the engine block on the engine mounts. This would be inefficient and frustrating, even with a highly skilled operator using a dual speed hoist. The high speed command would not be fast enough for moving from the factory floor to up and over the fender and the slow speed would not be slow enough for precise placement on the engine blocks. 
   Poor performance is tolerated in applications requiring super-human strength because there are few alternatives. However, there are countless applications in which a traditional hoist could be used to significantly reduce operator strain (for example placing a 50 pound car seat or 20 pound car battery) but is not used because of the frustration and inefficiency associated with the clumsy and slow operational features of traditional hoists. 
   SUMMARY OF INVENTION 
   According to one illustrative embodiment of the invention, there is provided an actuator for providing a force, having a base, a power source, a power transmission element coupled to the power source and constructed and arranged to move a load, and a physically compliant force measurement element constructed and arranged to provide at least partial support between the power transmission element and the base separate from the power source. The force measurement element deflects in relation to the force on the load. 
   In another embodiment, a hoist is provided, having a baseplate, a power source, a power transmission element coupled to the power source and constructed and arranged to at least partially support a load, and an elastic element that is coupled to the baseplate and supports at least a portion of the power transmission element on the baseplate without the elastic element exerting a substantial force on the power source. 
   In yet another illustrative embodiment of the invention, there is provided a method of controlling a hoist having a power source. The method comprises measuring the deflection of an elastic element that provides support between a hoist base and a payload without the support passing through the power source, providing a signal to a controller that indicates the deflection measurement, and providing a control signal that actuates the power source in relation to the measurement of the elastic element deflection. 
   Other features and aspects of the invention will be apparent from the detailed description, the figures, and the claims. 

   
     FIGURES 
       FIG. 1  is a perspective view of a hoist according to one illustrative embodiment of the invention; 
       FIG. 2  shows a drive train subassembly, an armature subassembly, and a spool subassembly in an exploded view; 
       FIG. 3  shows an example operational setup for a hoist; 
       FIG. 4  shows a side view of a hoist in an unloaded configuration according to an illustrative embodiment of the invention; 
       FIG. 5  shows a side view of the hoist shown in  FIG. 4  in a fully loaded configuration; 
       FIG. 6  shows forces acting on a spring element; 
       FIG. 7  is a flow chart of an example high-level control algorithm; 
       FIG. 8  is a block diagram of an example force feedback controller; 
       FIG. 9  is a block diagram of an example velocity feedback controller; and 
       FIG. 10  is a side view of a hoist according to another illustrative embodiment of the invention. 
   

   DETAILED DESCRIPTION 
   It would be desirable to be able to perform tasks such as lifting an engine block or a 50 pound car seat or a 20 pound car battery using a device that counterbalanced the payload weight and could operate over a continuously variable range of speeds. Input commands could flow directly from the user to the load instead of through a remotely controlled pendant. In this way, the operator could firmly grasp a payload such as an engine block with both hands and lift if off the factory floor and up and over the front fender at a natural speed, as if it weighed less than one pound. Once over the engine bay, the engine block could be slowly lowered into position and jostled into place almost effortlessly. 
   The industry has responded to this type of need with various payload counterbalancing devices, including pneumatic balancers, spring balancers, servomotor controlled balancers, and load cell balancers. The performance of these balancers is measured by their ability to:
         (a) manually counterbalance varying payloads. It is desirable for a counterbalancing mechanism to manually accommodate varying payloads easily because payloads can change weight from operation to operation.   (b) automatically counterbalance dynamically varying payloads. Additionally, the ability to automatically adjust to a dynamically varying payload is important when payload weight changes as an assembly process proceeds. To achieve dynamic counterbalancing, the counterbalancing device typically has a force measurement sensor which can sense change in weight.   (c) present the operator with a small ‘dead-band’. The balancing dead-band refers to the additional force required to move a counterbalanced payload up or down, thus a large dead-band requires significant operator effort even to move a counterbalanced payload. Large dead-bands are especially detrimental to performance if the dead-band is significant when compared to the payload. For example, a 10 pound dead band (i.e. 10 pounds of operator force) to move a 1000-pound payload may be acceptable, but the same 10-pound dead-band with a 20 pound payload may not be acceptable. Dead-bands are usually the result of static friction in the counter-balancing device.   (d) embody a small physical size. Reduced physical size is desirable in cramped manufacturing facilities. Typically, it is desirable to reduce the vertical dimension of the counterbalancing mechanism because it plays a role in the maximum floor clearance of the payload.       

   Spring balancers use constant force springs to counterbalance payloads. Spring balancers can manually accommodate payload changes if the operator manually adjusts spring tension. However, the counterbalancing force can be changed by only a small amount. 
   With traditional spring balancing hoists, it is difficult to dynamically change the counterbalancing force for dynamically varying payloads. With few moving parts, spring balancers inherently have very little friction and thus have very small dead-bands. For relatively small payloads, spring balancers can be designed to be physically compact. Unfortunately, the material properties of spring steel have prevented them from being successfully scaled to meet the need for heavier payloads. Given these performance characteristics, spring balancers are often used for cramped spaces where lightweight payloads change only occasionally. Spring balancers are less suitable for counterbalancing large and dynamically varying payloads. 
   Pneumatic balancers use air pressure inside pneumatic cylinders to provide counterbalancing force to payloads. Pneumatic balancers can be changed manually (i.e. pressing a button) with clever design of control relays that actuate pressure regulators. With no sensor to measure force, pneumatic balancers typically do not accommodate dynamically changing loads. Because of airtight seals in the pneumatic cylinder, the static friction in pneumatic balancers is high, resulting in a large dead-band. With such a large dead-band, pneumatics are less than desirable for small loads where the dead-band might be a significant percentage of the payload itself. Because each inch of travel adds an inch of length to the pneumatic cylinder, pneumatic balancers have the further problem that they can become cumbersome for large ranges of motions. Given these performance characteristics, pneumatic balancers are suited more for counterbalancing heavy and varying payloads, where sufficient overhead space can accommodate large height requirements. They are not as well suited for counterbalancing light payloads or payloads that require large ranges of vertical motion or vary dynamically. 
   An alternative to spring balancers and pneumatic balancers is a servomotor controlled balancer. By using a servomotor, the torque of the spool (thus the counterbalancing force on the load) can be accurately controlled using a well-known relation between motor torque and motor current. By turning a knob to control motor current, the user can manually balance varying payloads. With no sensor to measure force, servomotor balancers typically do not accommodate dynamically changing loads. Servomotors typically operate very inefficiently at low speeds and high torques, as is often the case when they are used in hoists. To compensate for poor efficiency, the servomotor would have to be considerably over-sized for use in a counterbalancing hoist, resulting in a cumbersome design. Alternately, a smaller motor could be operated very efficiently at high speeds and low torque. A gear reduction could be used to reduce the speed and increase the torque for this application. The use of a gear reduction may introduce significant friction and increase the reflected inertia at the output of the gearbox. In fact, friction can become essentially infinite in some types of non-backdriveable gear reductions with large reduction factors. Such a design would result in a large, or even infinite, dead-band. Given these performance characteristics, servomotor balancers are more suitable for balancing varying payloads if the size and expense of an oversized servomotor is not a concern. They are less suitable for counterbalancing payloads that vary dynamically. 
   The final category we discuss is load cell balancers. Wannasuphooprasit, et al. in U.S. Pat. No. 6,241,462 disclose a hoist that has a load cell which allows it to counterbalance dynamically varying payloads. The hoist actively controls the force on the load cell (thus the payload) through a feedback controller, where the actual load force is measured using a load cell and the motor is servoed to correct for differences between the desired load and actual load. This sensing and control scheme is commonly used to control force. With feedback control, small dead-bands are achievable. Since it is unnecessary to use excessively large motors or linearly actuated cylinders, physically compact designs are attainable. Given their performance characteristics, load cell balancers are often suitable for balancing both lightweight and heavy payloads that vary dynamically. 
   Load cells can be sensitive to shock loads and due to the high mechanical stiffness of load cells, controller gains are often kept relatively low to insure stability of the feedback control loop. Low control gains result in sluggish response times and non-optimized dead-bands. A further drawback is the presence of ‘chatter’, a phenomenon that is common in load cell systems when in contact with stiff environments. 
   Pratt et al., in U.S. Pat. No. 5,650,704, entitled “Elastic Actuator For Precise Force Control”, the entirety of which is hereby incorporated by reference, disclose a novel actuation scheme, dubbed “Series Elastic Actuation” in which an elastic element is intentionally placed in series between a motor and a load. Pratt et al. recognized that incorporating an elastic element in series with the payload allows the introduction of high control gains (relative to those achievable with load cell force control). As a result of high control gains, low impedance and high force fidelity were achieved. Additionally, the series elastic element provides inherent shock tolerance. Robinson describes these advantages in detail in Robinson, D. W. ‘Design and Analysis of Series Elasticity in Closed-loop Actuator Force Control’, Ph.D. Thesis, Massachusetts Institute of Technology, 2000, the entirety of which is hereby incorporated by reference. 
   Although Series Elastic Actuators show a marked improvement in performance as compared to typical force controlled actuators utilizing load cells, there remains a disadvantage: the actuator motion is bounded and typically small. This limitation is due to the need for the elastic element to move with the load. If the movement of the elastic element is linear, then the actuator&#39;s motion may be bounded by the stroke length of the actuator. If the movement of the elastic element is rotary, then the actuator&#39;s motion may be limited by sensor wires that measure force in the elastic element. In such an arrangement, the amount of rotation may be limited as the sensor wires may become overly twisted. In many applications, a limited motion is acceptable. For example, a joint in a robot arm or leg requires limited actuator motion since the joint can only rotate a fraction of a turn. In other applications, such as hoists and cranes, large motion may be required, and therefore Series Elastic Actuators, as disclosed in U.S. Pat. No. 5,650,704 may not be entirely suitable. 
   According to various embodiments disclosed herein, an actuator is presented for aiding in the lifting or moving of loads. In one embodiment, a spring-loaded counterbalancing hoist with improved dead-band and shock tolerance allows an operator to move a payload while the hoist dynamically counterbalances the payload weight. In some embodiments, a compliant element (e.g., a compression spring, a torsional spring, a rubber element, etc.) is combined with a position transducer (e.g., a potentiometer, a strain gauge, an optical encoder, etc.) to measure the force of the payload. Higher control gains, as compared to force control algorithms using load cells, allow gear reduction friction and motor inertia to be masked to a greater degree. Masked friction and inertia can result in a further reduction of the dead-band. In some embodiments, an actuator has a power source and a power transmission element and the compliant element at least partially supports the power transmission element. For purposes herein, a power transmission element can comprise some or all of the elements that transmit power from the power source output to the load. The power transmission element may include drive transmission assemblies, armature assemblies, gearboxes, pulleys, idle pulleys, cables, etc. Several aspects of various embodiments of the present invention with relation to conventional counterbalancing devices include:
         (a) The option to manually change the counterbalancing force to accommodate varying payloads. The counterbalancing force can be continuously variable (as opposed to discrete changes of pneumatic counterbalances) and can be realized with the push of a button (as opposed to spring balancers).   (b) The ability of the device to automatically change the counterbalancing force dynamically to accommodate varying payloads. Because spring, pneumatic and servomotor balancers do not have force-sensing elements, they do not have this ability. Compared to load cell balancers, a force-sensing elastic element is inexpensive and robust to shock loads.   (c) A small dead-band. An improvement in the dead-band, as compared to spring, pneumatic, and servomotor balancers, can be achieved with closed-loop feedback. High control gains, realizable with the use of an elastic element for force sensing may provide a better dead-band is than a load cell balancer.   (d) physically compact design realized with small motors and large gear reductions, feasible because of the use of an elastic element for force sensing.   (e) inherent shock tolerance due to the elastic element.       

     FIG. 1  is a perspective view of a spring-loaded counterbalancing hoist according to one illustrative embodiment of the invention.  FIG. 2  shows a partially exploded view of the hoist shown in  FIG. 1  with a drive train subassembly  58 , an armature subassembly  60  and a spool subassembly  62 . The drive train subassembly  58  includes a shaft encoder  20 , a brake  22  and a servomotor  24  concentrically aligned and affixed to one another. The mechanical output of servomotor  24  is coupled to the input of a gear reduction  26 . A drive shaft  32  is coupled to the output of gear reduction  26  by a keyway  28 . The drive shaft  32  is simply supported at its far end by a drive shaft support bearing  41 , which is mounted in a bearing housing  40 . A drive gear  38  is mounted on drive shaft  32  near bearing housing  40 . Gear reduction  26  and bearing housing  40  are mounted on a base, such as baseplate  30 . Other types of bases are contemplated, for example, a chassis, a robot link, and so on. A motor amplifier  34  and a controller  36  are also mounted on baseplate  30 . The armature subassembly  60 , shown in  FIG. 2 , includes a left armature  42   a  and a right armature  42   b  located on each side of the drive gear  38 . A left armature bearing  44   a  and right armature bearing  44   b  are mounted in left armature  42   a  and right armature  42   b , respectively. A spool shaft  50  connects armatures  42   a  and  42   b  to one another. A left compression spring  46   a  and a right compression spring  46   b  are affixed to armatures  42   a  and  42   b , respectively. The free ends of compression springs  46   a  and  46   b  rest on baseplate  30 . For purposes herein, “connected to” or “coupled to” do not require that two elements be physically attached. For example, compression springs  46   a  and  46   b  are connected and coupled to baseplate  30  even though the free ends of the springs may rest on baseplate  30 . A position transducer, such as a potentiometer  48 , is connected between left armature  42   a  and baseplate  30 . Of course, other types of position transducers may be employed, such as strain gauges, conventional hall effect sensors, magnetic position transducers, and optical position transducers, among others. Finally, the spool subassembly  62 , shown in  FIG. 2 , includes a spool gear  52  which meshes with drive gear  38  and spins freely on spool shaft  50  by way of a left spool bearing  56   a . A second spool bearing  56   b  is mounted concentrically in a spool  54 . Spool  54 , a left spool flange  53   a  and a right spool flange  53   b  are affixed concentrically to spool gear  52 . 
   Other compliant elements may be used in place of compressions springs  46   a  and  46   b . For example, torsional springs or rubber elements may be used. The compliant elements may be constructed of various suitable materials, for example, steel, aluminum, delrin, or nylon. In some embodiments, one compliant element along may be used. In other embodiments, two or more compliant elements may be used. If compression springs are used, such as in the embodiment shown in  FIG. 1 , each spring may have different properties. For example, compression spring  46   a  may be stiffer than compression spring  46   b.    
     FIG. 3  shows a typical mounting arrangement for the hoist. The hoist is enclosed in a hoist housing  74 . A mounting plate  72  protrudes from the top of the hoist housing  74  and is attached to a moveable overhead carriage  70 . A payload  80  is attached to a payload hook  78  at the end of payload cable  76  which is helically wound and terminated on spool  54 . A control pendant  86  is in communication with the controller  36  and the motor amplifier  34  via a communication cable  88 . The control pendant  86  includes an on/off button  90 , a float button  92 , an array of system status LEDs  94 , a down button  96 , and up button  98 , and a fast button  100 . 
   OPERATION 
   Operational Description— FIG. 3  shows an operational setup for the hoist according to one illustrative embodiment of the invention. Overhead carriage  70  allows the operator to move the hoist in two directions above the workspace  102 . Payload  80  is attached to the hoist via payload hook  78 . The operator commands the hoist to move payload  80  up and down in the vertical direction. In this embodiment, the operator has two modes of operation to choose from: float mode or manual up/down mode. Float mode is selected by depressing float button  92  on pendant  86 . In float mode, the weight of payload  80  is actively counterbalanced by the hoist with a closed-loop feedback control algorithm described below. Thus, the operator can apply an upward force  82  or a downward force  84  directly to payload  80  to move it in the desired vertical direction. The forces  82  and  84  may be small compared to the weight of a large load, allowing the operator to move the load easily and intuitively while expending less energy. Alternately, the operator may choose to operate in manual up/down mode. In manual up/down mode, the hoist performs like a traditional hoist. The operator issues velocity commands remotely from a control pendant  86 . If the user pushes up button  98 , the hoist will move the load upward at a moderate speed. If the user pushes down button  96 , the hoist will move the load downward at a moderate speed. If the fast button  100  is pressed while the up button  98  or down button  96  is also pressed, the hoist will move the load  80  up or down at a faster speed. 
   Mechanical Operation—Referring to  FIG. 1 , the following describes the motion of parts as the hoist is operated to lift a load. In lifting a load, servomotor  24  powers the gear reduction  26  causing the drive shaft  32  and attached drive gear  38  to rotate. Drive gear  38  meshes with spool gear  52 , thereby rotating spool  54 . Depending on the rotational direction of servomotor  24 , spool  54  winds or unwinds payload cable  76  (see  FIG. 3 ) and hence lowers or raises payload  80 . Brake  22  can be used to lock servomotor  24  in place, thereby preventing payload  80  from moving, except for small motions afforded by the compression of springs  46   a  and  46   b . Upon power-up the brake  22  is initially engaged. The brake  22  remains engaged until the operator issues a command via the control pendant  86 . Upon power down or power failure, the brake engages automatically via a spring-loaded mechanism to prevent the load from falling. A watchdog timer circuit may also be employed to lock the brake  22  in cases of controller  36  failure. 
   While the hoist shown in  FIG. 1  is described in connection with lifting and lowering loads, the components may be arranged to push and/or pull on objects. For example, the actuator in the hoist may be configured to power a robotic joint. 
     FIG. 4  shows a side view of the hoist in an unloaded configuration according to one illustrative embodiment.  FIG. 5  shows a side view of the hoist in a fully loaded configuration, with the load supported on baseplate  30  by compression spring  46 . A load, or other element, is considered “supported on” or “supported by” the base even if the load or other element is positioned below the base. Together,  FIG. 4  and  FIG. 5 . show how the force on payload  80  can be measured. As the force on payload cable  76  increases, springs  46   a  and  46   b  deflect to counteract a portion of the force. To understand this operation it is instructive to first examine how the springs deflect due to a load when the brake  22  is engaged, thereby fixing drive shaft  32  and drive gear  38 . Because spool gear  52  meshes with drive gear  38 , the spool gear  52  is unable to rotate freely about spool shaft  50  unless drive gear  38  is also able to rotate. This prevents load  80  from unwinding payload cable  76  freely from spool  54 . Still, even when drive gear  38  is locked in place, the armature subassembly  60  and spool subassembly  62  remain free to rotate with respect to drive gear  38  because of armature bearings  44   a  and  44   b  and the spool bearings  56   a  and  56   b . Thus, payload  80  produces a downward force on the armature subassembly  60 , causing the armature subassembly  60  to rotate CCW around drive shaft  32 . Springs  46   a  and  46   b  provide a counterbalancing force stopping the CCW rotation of armature subassembly  60 . 
   The force that springs  46   a  and  46   b  apply to counterbalance the load force can be computed using the free body diagram of  FIG. 6 . In the following calculations, we assume that the armature and spool subassemblies are in equilibrium such that the forces on the spool  54  sum to zero and the torques about the armature bearings  44   a  and  44   b  sum to zero. This assumption is valid since the mass of the spool subassembly  62  and armature subassembly  60  is small compared to typical loads. We also assume that the forces from the load, springs, and drive gear are all in the vertical direction. This approximation is valid since the angle the armatures  42   a  and  42   b  rotate is typically small. One could relax both of these assumptions to derive similar equations but we keep the assumptions here to avoid confusion. 
   There are three forces acting on the spool  54 . The load applies a downward force of F_load. The spring applies an upward force of F_spring. The drive gear supplies a downward force of F_drive_gear. Applying a force balance, we get
 
 F _load+ F _drive_gear= F _spring.  (1)
 
   There are three torques acting about armature bearings  44   a  and  44   b . The load applies a counterclockwise torque of F_load*(R_cable+2*R_gear) where R_cable is the distance from spool shaft  50  to the cable exit point from the spool shaft; R_gear is the radius of the drive gear  38  and spool gear  52 . The spring applies a clockwise torque of F_spring*R_spring where R_spring is the distance from the drive shaft  32  to the springs  46   a  and  46   b . The drive gear applies a counterclockwise torque of F_drive_gear*R_gear. Equating the sum of torques about the armature bearings  44   a  and  44   b  to zero, we get
 
 F _load*( R _cable+2 *R _gear)− F _spring* R _spring+ F _drive_gear* R _gear=0  (2)
 
   By algebraically manipulating Equations 1 and 2 to eliminate F_drive_gear, we can solve for F_spring as a function of F_load, or F_load as a function of F_spring:
 
 F _spring=( R _gear+ R _cable)/( R _spring− R _gear)* F _load  (3)
 
 F _load=( R _spring− R _gear)/( R _gear+ R _cable)* F _spring  (4)
 
   The force on the spring can be calculated using Hooke&#39;s Law (F=Kx) where K is the known spring constant of compression springs  46   a  and  46   b  and x, the deflection of the spring or springs, is measured with a deflection measurement device, such as a potentiometer  48 . For a non-linear spring, a similar relation can be used. F_load can then be computed using Equation 4. 
   Control System Operation—Referring to the illustrative embodiment shown in  FIG. 3 , the control pendant  86  sends the operator&#39;s commands to controller  36  via the communications cable  88 . The controller  36  accepts signals from the control pendant  86 , potentiometer  48 , and shaft encoder  20 , and sends commands to motor amplifier  34 , which sends electrical current to servomotor  24 . 
     FIG. 7  shows a high-level flow chart of one illustrative embodiment of a control algorithm that may be executed by controller  36 . This embodiment is presented as an example as many other suitable algorithms may be used. Example values are shown in  FIG. 7 , but as should be evident to one skilled in the art, any suitable values may be used. On power-up, the algorithm starts at step  1000 , sets the desired velocity to zero, engages the brake, disables the motor amp, and sets the mode to manual up/down. Step  1010  is then entered. Since the hoist starts in manual up/down mode, the controller moves to step  1020 . If neither the up button or the down button is pushed, then the desired velocity starts ramping to zero in step  1210 . Step  1220  checks if the desired velocity is zero and if so, engage the brake in step  1230  and disables the motor amp in step  1240 . If the desired velocity is not zero in step  1220 , then the brake is disengaged in step  1190  and the load is servoed to the desired velocity in step  1200 . ( FIG. 9  shows the block diagram for the velocity controller which is explained below.) If either the up button or the down button is pushed, then step  1020  or step  1030  detects it and the desired velocity ramps toward a desired speed in one of steps  1150 ,  1160 ,  1170 , or  1180 , depending on which button was pressed and whether the fast button was pushed (steps  1130  and  1140 ). After determining the desired velocity, the brake is disengaged in step  1190  and the desired velocity is servoed in step  1200 . Whether servoing the desired velocity in step  1200  or disabling the motor amp in step  1240 , the controller next enters step  1250  and checks if the up button or down button is pushed. If so, then the mode is set to manual up/down in step  1260  and the controller loops back to step  1010 . 
   If the up button and down button are not pushed in step  1250 , then the controller enters step  1270  and checks if the float button is pushed. If not, then the mode is set to manual up/down in step  1260  and the controller loops back to step  1010 . If the float button is pushed in step  1270 , then the weight of the load is estimated, sampled, and set as the desired force in step  1280 . The controller sets the mode to “float” in step  1290  and loops back to step  1010 . In step  1010 , if the mode is set to “float”, then the controller moves to step  1040  and determines if the hoist is idle (i.e., the load has not moved for a few seconds). If the hoist is idle, then the brake is engaged in step  1050  and the motor amplifier is disabled in step  1060 . If the hoist was not idle in step  1040 , then the brake is disengaged in step  1070  and the motor is driven in order to compress the spring to the desired force corresponding to the load weight measured in step  1280 . ( FIG. 8  shows the block diagram for the force controller, which is explained below.) Regardless of whether the hoist is idle, the controller moves to step  1090  from step  1080  or step  1060  and checks if the up button or the down button is pushed. If so, the mode is set to “manual up/down” in step  1100  and the controller loops to step  1010 . If not, then the controller moves to step  1110  and checks if the float button is pushed. If the float button is not pushed, the mode is set to manual up/down in step  1100  and the controller loops to step  1010 . If the float button is pushed, then the mode is set to “float” in step  1120  and the controller loops to step  1010 . Of course, any suitable order of operations or control sequences may be used to control the operation of the actuator or hoist components, and the above flow chart and description is provided by way of example only. 
   With this control algorithm, to move the payload  80 , the operator may first select a mode of operation by depressing the float button  92 , the manual up button  98 , or the manual down button  96  on the control pendant  86 . If the manual up or manual down buttons are pressed, the hoist behaves like a traditional velocity controlled up/down hoist. If the float button is pushed, then the hoist suspends the load by applying an upward force on the load that counteracts gravity. The user can then move the load up or down by manually applying a force to the load that is much smaller than the weight of the load. Thus, the load feels virtually weightless to the operator in float mode. 
     FIG. 8  shows a block diagram of an example force control servo that may be run in step  1080  of  FIG. 7 . This force control servo is a standard Proportional-Derivative (PD) control loop that servos to the actual force (i.e., the counterbalancing force) to match the desired force. In block  2080 , the spring deflection is measured and converted to the actual spring force in block  2070 . This force is subtracted from the desired spring force in block  2000  to get the force error. This error is multiplied by a proportional gain, K, in block  2010  and added to the derivative of the error (block  2020 ) times a derivative gain, B (block  2030 ), in block  2040 . The resultant signal then goes to motor current amplifier in block  2050 , which then drives the servomotor  24 . Other method or control algorithms for actuating the power source in relation to the measured spring deflection may be used. For purposes herein, “in relation to” a quantity (such as spring deflection) does not imply in relation to only that quantity. The power source may be actuated, or other actions may be taken, in relation to other inputs or control signals. 
   Referring to  FIG. 1 , we illustrate how the controller may interact with the hardware to control the force on the load. If the actual force is greater than the desired force, current is sent to servomotor  24 , which causes the drive gear  38  to rotate CW and spool gear  52  to rotate CCW. As a result, spool  54  unwinds cable  76 , accelerating load  80  downward, which has the effect of dynamically decreasing the actual force exerted on the compression springs  46   a  and  46   b . Conversely, if the actual force is less than the desired force, then current is sent to servomotor  24  causing drive gear  38  to rotate CCW and spool gear  52  to rotate CW. As a result, spool  54  winds cable  76 , accelerating load  80  upward, which has the effect of dynamically increasing the actual force exerted on the compression springs  46   a  and  46   b . This is an example of one manner in which the controller can correct for differences between the desired and actual forces on the load. 
     FIG. 9  shows a block diagram of an example velocity control servo that may be run in step  1200  of  FIG. 7 . In block  2180 , the motor velocity is measured from the shaft encoder on the back of the motor. This measurement is then converted to load velocity in block  2150 . In block  2190 , the spring deflection is measured and differentiated in block  2170 . This signal is then converted to load velocity in block  2160 . The measurements from block  2150  and block  2160  are then added in block  2200  to get the actual load velocity. The load velocity is then subtracted from the desired velocity in block  2100  to get the velocity error. The velocity error is multiplied by a gain G in block  2110  and added to the integral of the error (block  2120 ) times an integral gain (block  2130 ) in block  2140 . The resultant signal is a desired force that is sent to the force controller in  FIG. 8 , the operation of which is described above. Thus, if there is an error in the velocity of the load, a force will be exerted on the load to correct for the velocity. 
   Other embodiments of this invention are envisioned. For example, in another embodiment shown in  FIG. 10 , a hoist similar to the one shown in  FIG. 1  is provided. A power source  102 , connected to a base  101 , actuates cable  103 . Cable  103  winds over idle pulley  105  and then attaches to load  104 . Idle pulley  105  is at least partially supported by elastic element  106 . The deflection of elastic element  106  relates to the force applied by cable  103  on load  104 . 
   While the above description has been discussed with relation to counterbalancing hoists, various aspects of the embodiments may be used for other applications such as, for example, actuators, hoists, robots, elevators, and industrial machinery. 
   In view of the wide variety of embodiments to which the principles of the invention can be applied, it should be understood that the illustrated embodiments are exemplary only, and should not be taken as limiting the scope of the present invention. In addition, certain aspects of the present invention can be practiced with software, hardware, or a combination thereof.

Technology Classification (CPC): 1