Patent Publication Number: US-11047458-B1

Title: Load-adjustable constant-force mechanisms

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This nonprovisional application is a continuation of and claims priority to, U.S. Non-Provisional patent application Ser. No. 15/679,245, entitled “Load-Adjustable Constant Force Mechanisms,” filed Aug. 17, 2017 by the same inventors, which claims priority to U.S. Provisional Patent Application No. 62/376,177, entitled “Load-Adjustable Constant-Force Mechanisms”, filed Aug. 17, 2016 by the same inventors, the entirety of which is incorporated herein by reference. 
    
    
     FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT 
     This invention was made with Government support under Grant No. CMMI-1000138 awarded by the National Science Foundation. The government has certain rights in the invention. 
    
    
     BACKGROUND OF THE INVENTION 
     1. Field of the Invention 
     This invention relates to compliant mechanisms. More specifically, it relates to compliant constant-force mechanisms. 
     2. Brief Description of the Related Art 
     Compliant Mechanisms 
     A compliant mechanism is a flexible mechanism that derives some or all its motion (mobility) from the deflection of flexible segments, thereby replacing the need for mechanical joints. It transfers an input force or displacement from one point to another through elastic body deformation. The absence or reduction of mechanical joints impacts both performance and cost. Advantages include reduced friction and wear, increased reliability and precision, and decreased maintenance and weight. Moreover, the cost is also affected by reduced assembly time and, in most cases, due to its hingeless design, the fabrication of such mechanisms can be produced from a single piece. Additionally, compliant mechanisms provide the designer with an effective way to achieve mechanical stability. 
     These mechanisms are most commonly designed in two ways. One is using pseudo-rigid-body models, and the other is using topology optimization. Both approaches have utility. The design of the compliant portion of the unit cell components is accomplished through compliant mechanism synthesis. 
     There are three major approaches to the design and synthesis of compliant mechanisms: kinematic approximation methods, computationally intense methods, and linear and higher-order expansions of the governing equations. This disclosure is based primarily upon kinematic approximation methods. 
     The kinematic approximation or Pseudo-Rigid-Body Model (PRBM) approach works by identifying similarities between compliant mechanisms and rigid-body mechanisms. It has proved effective in identifying numerous compliant analogs to ubiquitous planar rigid-body mechanisms such as four-bar and crank-slider mechanisms. The chief criticisms of this approach are that the models are approximate and have limited, albeit known, accuracy. Moreover, the identification between flexure geometries and rigid-body mechanisms has been limited to a small but versatile set of planar configurations. 
     Computationally intense approaches typically combine finite element analysis with optimization to calculate optimal geometries in response to load and motion specifications. This approach has been successful but has also been criticized for producing results identical to those produced more quickly by the PRBM approach, or results that are not physically realizable. As a general rule, this approach is more capable and accurate than the PRBM approach, but also more time-consuming. 
     The third approach, which relies on linear and higher-order expansions of the governing equations, is well-known in precision mechanisms research and relies heavily on flexures that are small and undergo small, nearly linear, deflections. This approach uses flexures much smaller than the overall mechanism size, so it is not generally applicable to millimeter-scale and smaller mechanisms. These techniques are important but do not have a direct bearing on the invention disclosed herein. 
     Constant-Force Mechanisms 
     Constant-force mechanisms (CFM) already exist in the form of springs, rigid-body mechanisms, and in the form of compliant mechanisms (i.e., flexible devices which transmit force or displacement by means of elastic deformation). An example of use of CFMs is in robotic end effectors. A typical robotic arm has no real way of controlling the amount of contact pressure it applies. By adding a constant-force end effector, the arm only needs to operate within a certain range of motion to maintain the desired force output. For example, the result can be observed in a glass-cutting robot. With a CFM in place, its end-effector could have a high positional error allowance and still apply the correct amount of force to cut the glass without breaking it. The limitation is that these devices must be designed around their intended force output, and as such, an issue arises with how the same robot would be able to cut several thicknesses of glass. 
     In the conventional art, constant-force compliant mechanisms with constant stiffness are mechanisms that allow for substantial decoupling of force and linear displacement in robotic systems. In other words, the same force is transmitted regardless of the position of the mechanism; this is quite restrictive. Accordingly, what is needed is a system capable of varying the stiffness of these compliant CFMs. However, in view of the art considered as a whole at the time the present invention was made, it was not obvious to those of ordinary skill in the field of this invention how the shortcomings of the prior art could be overcome. 
     While certain aspects of conventional technologies have been discussed to facilitate disclosure of the invention, Applicants in no way disclaim these technical aspects, and it is contemplated that the claimed invention may encompass one or more of the conventional technical aspects discussed herein. 
     The present invention may address one or more of the problems and deficiencies of the prior art discussed above. However, it is contemplated that the invention may prove useful in addressing other problems and deficiencies in a number of technical areas. Therefore, the claimed invention should not necessarily be construed as limited to addressing any of the particular problems or deficiencies discussed herein. 
     In this specification, where a document, act or item of knowledge is referred to or discussed, this reference or discussion is not an admission that the document, act or item of knowledge or any combination thereof was at the priority date, publicly available, known to the public, part of common general knowledge, or otherwise constitutes prior art under the applicable statutory provisions; or is known to be relevant to an attempt to solve any problem with which this specification is concerned. 
     BRIEF SUMMARY OF THE INVENTION 
     The long-standing but heretofore unfulfilled need for a load-adjustable constant-force mechanism is now met by a new, useful, and nonobvious invention. 
     In an embodiment, the current invention is a constant-force mechanism that has a second degree of freedom that adjusts the output force without changing a kinematic structure of the mechanism. This second degree of freedom is the rotation of a beam about a longitudinal axis thereof constrained to an initial plane of bending. 
     In a separate embodiment, the current invention is a compliant, load-adjustable constant-force crank-slider mechanism. The mechanism includes a compliant beam (e.g., cross-section having an aspect ratio of about 2:1) formed of a first link and a second link, both of which are optionally flexible, where the first link extends away from a base. Each link has two ends. One end of the first link is linearly fixed. A torsional spring is disposed at the fixed end of the first link, such that the first link is rotatable about the spring. The torsional spring is biased toward a stable position and upon being unstable, outputs a constant force to return to the stable position. For example, if this stable position is a state of expansion when the mechanism undergoes compression, the constant force is outputted to expand the mechanism. In an embodiment, an angle of elastic deflection of the first link may increase as the mechanism compresses, and vice versa. 
     The non-fixed end of the first link and an end of the second link are coupled together by a joint (e.g., pin connection) disposed therebetween at a predetermined distance away from the base. The links are rotatable relative to each other about the joint. A slider is positioned at the opposite end of the second link. The slider traverses along an x-axis during compression and expansion of the spring. The beam is rotatable (e.g., freely rotatable) about its longitudinal axis, such that beam rotation alters the constant force that the spring outputs to return to its stable position. Optionally, to actuate beam rotation, a stepper motor or a compliant ratchet system may be used. In any case, even with this beam rotation, the crank-slider mechanism would be constrained to its initial axis of bending. 
     In a separate embodiment, the current invention is a compliant, load-adjustable constant-force crank-slider mechanism, including any one or more—or even all—of the foregoing characteristics and features. 
     It is an object of the current invention to perform the same function of a conventional CFM but also allows for an adjustable output. This is accomplished by rotating the flexible member of a compliant constant-force slider about its longitudinal axis. By constraining the compliant beam to always bend in the same plane, its stiffness changes as a result of its rotation. This allows the force output of the device to be adjusted as a function of the rotation of the compliant beam. 
     The present invention provides precise force output control independent of position. The robot-environment relationship requires force control for normal behavior and unexpected collisions. Stiff actuators are effective for precise positional control but are ineffective for force control. Therefore, it is favorable that the elasticity of this system transforms a problem of force control into one of positional control. 
     Closing the control loop with positional feedback also enables dynamic adjustability. CFMs are zero-stiffness devices, but dynamic load adjustment as a function of position would allow for tuning of stiffness up or down from zero. By adding positional feedback control to a system with innate force control, position can be set and maintained as a function of applied load. The substantial decoupling of force and displacement is beneficial because it means that the stiffness of the mechanism is subject to digital control. 
     These and other important objects, advantages, and features of the invention will become clear as this disclosure proceeds. 
     The invention accordingly comprises the features of construction, combination of elements, and arrangement of parts that will be exemplified in the disclosure set forth hereinafter and the scope of the invention will be indicated in the claims. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
       For a fuller understanding of the invention, reference should be made to the following detailed disclosure, taken in connection with the accompanying drawings, in which: 
         FIG. 1  depicts an embodiment of a compliant constant-force slider and its equivalent pseudo-rigid body model. 
         FIG. 2  depicts a pseudo-rigid body model of the compliant crank slider depicted in  FIG. 1 . 
         FIG. 3  is a graph depicting a theoretical force output bandwidth. 
         FIG. 4A  is a perspective view of compliant CFM, according to an embodiment of the current invention that that uses a compliant ratchet system to enact beam rotation. 
         FIG. 4B  is a side view of the compliant CFM of  FIG. 4A . 
         FIG. 5  is a photograph depicting a prototype created in accordance with an embodiment of the invention 
         FIG. 6  is a graph depicting recorded force-displacement curves compared to theoretical constant-force performance at several rotational settings. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
     In the following detailed description of the preferred embodiments, reference is made to the accompanying drawings, which form a part thereof, and within which are shown by way of illustration specific embodiments by which the invention may be practiced. It is to be understood that other embodiments may be utilized, and structural changes may be made without departing from the scope of the invention. 
     As used in this specification and the appended claims, the singular forms “a”, “an”, and “the” include plural referents unless the content clearly dictates otherwise. As used in this specification and the appended claims, the term “or” is generally employed in its sense including “and/or” unless the context clearly dictates otherwise. 
     As used herein, “about” means approximately or nearly and in the context of a numerical value or range set forth means±15% of the numerical. In an embodiment, the term “about” can include traditional rounding according to significant figures of the numerical value. In addition, the phrase “about ‘x’ to ‘y’” includes “about ‘x’ to about ‘y’”. 
     A beam&#39;s stiffness is its resistance to bending deformation. In the case of compliant mechanisms, this bending is idealized as purely elastic and is a function of two properties—elastic modulus and second moment of area. To demonstrate, CFM in its simplest form exists in a compliant crank slider. It implements link proportions that produce a roughly constant-force output. This was achieved by substituting one or more compliant beams into a traditional crank slider. The particular configuration that the present invention emulates is depicted in  FIG. 1 . 
     Pseudo-Rigid Body Model (PRBM) 
     The PRBM treats a beam in pure bending as two rigid links with a torsional spring at their joint/pin connection. A fixed-free beam bends very little at the end furthest away from its base. As shown in  FIG. 1 , base  12  is modeled as an extension of the ground, and the deflecting portion of beam  14  is modeled as a rigid link  16  that is pinned to the base  12 . This results in a circular path for beam tip  18 , which is not exactly accurate, but it is close enough to be a viable approach. A small amount of error may be acceptable to simplify the analysis of compliant mechanisms  10 . 
     A new parameter is defined for this model, specifically how far along beam  14  to position pin connection  20 . This length is defined by γ. A standard value of γ=0.83 signifies that 83% of compliant beam&#39;s  14  length is treated as rotating about a fixed point that is located 17% of the beam&#39;s length away from its fixed boundary. 
     The input motion of the above compliant slider mechanism  10  is assumed to be slider  22  deflection, Δx, so that r 1 =r 2 +r 3 −Δx. As mechanism  10  is compressed, the angle of elastic deflection θ 2  will increase. 
     Torsional spring  24  k 1  describes the energy stored in beam  14  as it is deformed elastically, where: 
     
       
         
           
             
               
                 
                   
                     
                       k 
                       1 
                     
                     = 
                     
                       γ 
                       ⁢ 
                       
                         K 
                         θ 
                       
                       ⁢ 
                       
                         
                           EI 
                           ⁢ 
                           
                               
                           
                           ⁢ 
                           
                             θ 
                             2 
                           
                         
                         l 
                       
                     
                   
                   . 
                 
               
               
                 
                   ( 
                   1 
                   ) 
                 
               
             
           
         
       
     
     Each term in Equation 1 has been discussed except for K θ , which is a non-dimensional stiffness coefficient approximated as γπ. The output of mechanism  10  is the reaction force on slider  22 , which is assumed to move in line with the fixed end of compliant beam  14 . 
     Virtual Work 
     Real-world application of this model considers partial weight and center of gravity. 
     Virtual work analysis accounts for only the components of force and moment that affect the ultimate output. 
     In analyzing each contribution to the total force output of mechanism  10 , two orientations were defined for the device—horizontal orientation and vertical orientation. Changing the direction of gravity could affect behavior enough to require two separate models. 
     In each model, the mass of the pin connection between r 2  and r 3  dominates any divergence from theoretical performance. Minimizing the mass of this component can be difficult considering it is not a simple pin connection  20 , as compliant beam  14  should also be able to rotate freely about its longitudinal axis. As depicted in  FIG. 2 , W 3  is established as the weight of this component, and M 3  is the moment created by any center-of-mass offset from pin connection  20 . The virtual work model can then be solved for the force output as a function of beam properties, link dimensions, and angle of deflection. 
     For the horizontal model, 
     
       
         
           
             
               
                 
                   
                     F 
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                             2 
                           
                         
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                       - 
                       
                         
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                           2 
                         
                       
                       - 
                       
                         
                           M 
                           3 
                         
                         
                           r 
                           3 
                         
                       
                     
                     
                       
                         sin 
                         ⁢ 
                         
                           θ 
                           2 
                         
                       
                       - 
                       
                         sin 
                         ⁢ 
                         
                           θ 
                           3 
                         
                       
                     
                   
                 
               
               
                 
                   ( 
                   2 
                   ) 
                 
               
             
           
         
       
     
     For the vertical model, 
     
       
         
           
             
               
                 
                   
                     F 
                     out 
                   
                   = 
                   
                     
                       
                         
                           W 
                           3 
                         
                         ⁢ 
                         cos 
                         ⁢ 
                         
                           θ 
                           2 
                         
                         ⁢ 
                         sin 
                         ⁢ 
                         
                           θ 
                           2 
                         
                       
                       - 
                       
                         
                           ( 
                           
                             
                               γ 
                               2 
                             
                             ⁢ 
                             
                               K 
                               θ 
                             
                             ⁢ 
                             
                               EI 
                               
                                 r 
                                 2 
                                 2 
                               
                             
                             ⁢ 
                             
                               θ 
                               2 
                             
                           
                           ) 
                         
                         ⁢ 
                         cos 
                         ⁢ 
                         
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                           3 
                         
                       
                       + 
                       
                         
                           
                             M 
                             3 
                           
                           
                             r 
                             3 
                           
                         
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                         cos 
                         ⁢ 
                         
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                           2 
                         
                       
                     
                     
                       
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                         sin 
                         ⁢ 
                         
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                         cos 
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                         sin 
                         ⁢ 
                         
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                           2 
                         
                       
                     
                   
                 
               
               
                 
                   ( 
                   3 
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     W 3  and M 3  respectively are pin connection&#39;s  20  weight and moment contribution due to offset center of gravity. These models were produced using virtual work analysis. They provide roughly constant-force outputs as device  10  is compressed. 
     Beam&#39;s  14  bending stiffness about a non-principal stiffness axis can be calculated according to Mohr&#39;s circle. For beam  14  undergoing a rotation Γ about its longitudinal axis, the following values do not change. 
     
       
         
           
             
               
                 
                   
                     I 
                     _ 
                   
                   = 
                   
                     
                       
                         I 
                         
                           y 
                           ⁢ 
                           y 
                         
                       
                       + 
                       
                         I 
                         zz 
                       
                     
                     2 
                   
                 
               
               
                 
                   ( 
                   4 
                   ) 
                 
               
             
             
               
                 
                   
                     I 
                     R 
                   
                   = 
                   
                     
                       
                         I 
                         
                           y 
                           ⁢ 
                           y 
                         
                       
                       - 
                       
                         I 
                         zz 
                       
                     
                     2 
                   
                 
               
               
                 
                   ( 
                   5 
                   ) 
                 
               
             
           
         
       
     
     With beam  14  always constrained to bend about the Z-axis, the new second moment of area following beam  14  rotation can be calculated as follows:
 
 I   zz   ′=Ī−I   R  cos(2Γ)  (6)
 
     Equations 2 and 3 relate the force output of device  10 . The variables are second moment of area I, and joint angles θ 2  and θ 3 . Changes in second moment of area are a function of beam  14  rotation Γ and aspect ratio. 
     In certain embodiments, the current invention maximizes constant-force properties as a function of link  26  length from 0°≤θ 2 ≤40°. Pseudo-rigid link  26 A and  26 B lengths were selected to minimize the standard deviation of force outputs while maintaining the desired average output. Theoretical force output bandwidth is depicted in  FIG. 3 . 
     The present invention offers a theoretical 400% increase in force output through 90 degrees of beam  14  rotation. This is the result of beam&#39;s  14  cross-section having a 2:1 aspect ratio. It is possible to increase force bandwidth with a higher beam  14  aspect ratio. 
     A mechanical prototype relying on a compliant ratchet system  28  to enact beam  14  rotation is depicted in  FIGS. 4A-4B . Device  10  utilizes rigid links  16 A-D to hold compliant beam&#39;s  14  direction of bending constant. First rigid links  16 A and  16 B are pinned at a point described by the PRBM. If unconstrained, the fixed-free beam tip  18  would follow approximately the same circular path as rigid links  16 A-D. 
     By increasing or decreasing the stiffness of compliant mechanism  10 , mechanism&#39;s  10  resistance to linear motion can be actively adjusted. This is accomplished by rotating one or more flexible members along their longitudinal axis while constraining them to their initial axis/plane of bending. This corresponds to a change in second moment of area that increases or decreases the stiffness of mechanism  10 . In the case of compliant constant-force devices  10 , this rotation relates directly to a change in output force. Flexural pivots can be rotated in the same way to keep a system statically balanced in various positions. Compliant mechanism  10  designs utilizing beam  14  rotation in this manner allow for adjustment to a wide range of force levels, and also for continuous adjustment under computer control. 
     It is possible to adjust compliant mechanisms  10  in a way that directly affects their output force. This adjustment can be made either mechanically or electronically. Several applications present themselves. For example, constant-force devices, such as those used in robotic end effectors, become highly adjustable with the option of improving constant-force properties under computer control. Robotic armatures could also employ this system in their joints to maintain static balance in a wide range of positions, improving power efficiency. 
     EXAMPLE 
     While a purely mechanical implementation of beam  14  rotation maintains simplicity, computerized control allows for precise adjustment along a continuous range of force outputs. An exemplary design depicted in  FIG. 5  uses a stepper motor to enact beam rotation at the fixed base  12 . It is assumed that the unconstrained beam tip  18  follows rotation Γ at the fixed base  12 . 
     The bulk of the apparatus houses a NEMA  17  stepper motor with a 99.05:1 planetary gearbox. This motor setup relates a 0.01817° step interval for high-resolution adjustment of beam rotation Γ. 
     The results provided in  FIG. 6  were obtained with displacement-controlled force measurement. A tensile testing machine was fitted with a 50-pound strain gauge load cell and a zero-balance error of 1%. Each rotational adjustment produced a corresponding force increase in line with the theoretical model. Ultimately, zero-balance error correction of 1.8 pounds was applied uniformly to the data shown in  FIG. 6 . 
     Using materials having a higher stiffness would yield results closely aligned with those obtained through the theoretical model. For example, a steel compliant beam  14  would allow for a much higher aspect ratio and avoid issues arising from stress relaxation, creep, and plastic deformation. A higher beam  14  aspect ratio would increase force bandwidth and take full advantage of a high-resolution actuator. 
     Glossary of Claim Terms 
     Compliant: This term is used herein to refer a flexible mechanism transferring an input motion, energy, force, or displacement to another point in the mechanism via elastic body deformation. A compliant mechanism gains at least a portion of its mobility through deflection of its flexible components. 
     Crank-slider mechanism: This term is used herein to refer to a system of mechanical parts working together to transition between linear motion and rotating motion. 
     Initial axis/plane of bending: This term is used herein to refer to the plane within which the crank-slider mechanism undergoes deformation. When the beam is rotated about its longitudinal axis, the beam still deforms/bends in the same plane. 
     Joint: This term is used herein to refer to the point or structure where two components of a system join but are still able to rotate relative to each other. 
     Linearly fixed: This term is used herein to refer to a stationary position of a component relative to moving in the x-, y-, and z-axes. For example, an end of a link can be fixed in place, while the remainder of the link can rotate about that fixed point and/or rotate along its longitudinal axis. That fixed end, however, does not move. 
     Load-adjustable: This term is used herein to refer to the ability to alter a capacity of slider to be used in the underlying crank-slider mechanism. This is accomplished by being able to alter the constant output force exerted within the system. 
     Slider: This term is used herein to refer to the prismatic joint (e.g., piston) that undergoes linear movement as a result of actuation of the crank in the underlying crank-slider mechanism. 
     X-axis: This term is used herein to refer to the horizontal axis of a system of coordinates. However, it can be understood and is contemplated herein that this is a relative term dependent on the orientation of the system. The term “x-axis” is used to indicate the horizontal axis as depicted in  FIGS. 1-2 . However, if the system is angled in any way, the “x-axis” can be angled in the same way, for example, to indicate a similar path of travel of the slider. 
     Y-axis: This term is used herein to refer to the vertical axis of a system of coordinates. However, it can be understood and is contemplated herein that this is a relative term dependent on the orientation of the system. The term “y-axis” is used to indicate the vertical axis as depicted in  FIGS. 1-2 . However, if the system is angled in any way, the “y-axis” can be angled in the same way, for example, to indicate a similar path of travel of the joint. 
     The advantages set forth above, and those made apparent from the foregoing description, are efficiently attained. Since certain changes may be made in the above construction without departing from the scope of the invention, it is intended that all matters contained in the foregoing description or shown in the accompanying drawings shall be interpreted as illustrative and not in a limiting sense. 
     It is also to be understood that the following claims are intended to cover all of the generic and specific features of the invention herein described, and all statements of the scope of the invention that, as a matter of language, might be said to fall therebetween.