Abstract:
A force sensor apparatus includes a vibrating beam and first and second isolator mass members that supports ends of the vibrating beam. The first and second isolator mass members are configured symmetrically relative to an axis that intersects the vibrating beam at an angle other than 90 degrees. First and second end mounts connect respectively to the first and second isolator mass members. Each isolator mass member has a center of gravity. Each isolator mass member is shaped so that it can be massive (e.g., along the x-axis direction) while at the same time having its center of gravity at an optimal location so that undesirable beam forces and moments that would otherwise transfer vibrating beam energy to the end mounts are cancelled.

Description:
FIELD OF THE INVENTION  
         [0001]    The present invention relates to vibrating beam force sensors, and more particularly, to isolator mechanisms for isolating the vibrations of a vibrating member from its mounts to minimize coupling between the member and its mounts over a range of frequencies of vibration.  
         BACKGROUND AND SUMMARY OF THE INVENTION  
         [0002]    Vibrating beam force transducers find application in the instrumentation field in accelerometers, pressure transducers, scales, etc. In a vibrating beam force transducer, a vibratory beam member is supported so that axial forces to its ends change its axial stress in response to an input acceleration, pressure, etc. to be measured. In an axially unstressed condition, a beam has a particular natural frequency of vibration determined primarily by its dimensions, its material, its end conditions, and to a smaller extent the temperature and the media in which it is operating. As an applied force changes the axial tension or compression load on the beam, the beam&#39;s natural frequency of vibration changes. The frequency of its flexure vibration increases with axial tension and decreases with axial compression.  
           [0003]    Frequency modulated vibratory sensors are attractive in instrumentation because of the inherent high resolution, digital nature of the output signal. When the sensor material is quartz crystal, the sensor has excellent stability of bias frequency and span as well as low temperature sensitivity. The piezoelectric property of quartz crystal provides a simple way of sustaining the vibration using an oscillator circuit electrically connected to electrodes plated on the crystal beam resonator.  
           [0004]    Although it is desirable to have the vibration frequency output be a true and accurate representation of the actual force applied to the vibrating beam, this is not always the case. In practical applications, a vibrating beam force transducer is mounted to one or more end mounts. There is an energy loss at the mount interface because the mount resists the forces and moments generated by the vibrating beam resonator. This results in a decrease in the “Q” factor of the beam resonator, i.e., the ratio of peak energy to the energy lost per cycle of the vibrating beam resonator. In addition to decreased efficiency, a decrease in Q also degrades the frequency stability of the resonator. Thus, to achieve an efficient and stable vibrating beam resonator, it is desirable to design the transducer so that very little of the vibration energy is lost during operation.  
           [0005]    [0005]FIG. 1 illustrates a known vibrating beam force transducer apparatus  10  with a vibration isolation mechanism. A pair of end mounts  12  and  14  are attached to isolation members  18  and  20  by way of corresponding pairs of thin spring members  22 ,  24 , and  26 ,  28 , respectively. Each isolation member and corresponding spring member is called an isolator mass structure. Axial forces are applied along an X-axis to one or both end mounts  12  and  14  when the apparatus  10  is used as a force measuring unit such as in an accelerometer application. A vibrating beam  16  extends between the two isolation members  18  and  20 . The vibrating beam  16  is isolated from the mounts  12  and  14  at beam operating frequencies by the isolation members  18  and  20  as well as the thin spring members  22 - 28 . Axial stresses, either tension or compression, applied to the end mounts  12  and  14  are transported to the beam  16  through the thin spring members  22 ,  24 ,  26 , and  28  and isolator members  18  and  20 .  
           [0006]    Electrodes are used to drive the beam  16  so that it vibrates at a particular frequency. One pair of electrodes  38  and  40  is attached to opposite sides of the beam  16  along one axial extent, and another pair of electrodes  42  and  44  is attached to opposite sides of the beam along another axial extent. An oscillator circuit (not shown) provides driving excitation for the beam  16 . The oscillator circuit applies oppositely-directed, transverse electric fields at axially-spaced locations to vibrate the beam. This arrangement is a shear mode piezoelectric drive known in the art.  
           [0007]    The vibrating beam  16  in a momentary posture shown in FIG. 2 depicts the force-frequency effect of a vibrating flexure beam. The deflection is exaggerated for better illustration. The variable L corresponds to the length of the beam, t represents the thickness of the beam, b represents the width of the beam, and F represents the axial force on the beam.  
           [0008]    In the exploded drawing shown in FIG. 3, the various reactions to the beam  16  vibrating with amplitude Y B  are represented by reactive forces F 1Y , F 1X  and V and the reactive moment M. The V and M reactions are by far the largest. Also, because the vibration frequency of the beam  16  is much greater than the natural frequency of the isolation mechanism to both Y R  linear and A angular vibration modes, the phasing of the various linear and angular displacements is as shown in FIG. 3. When the beam is deflected at its root locations intermediate to its ends, there is a reaction force V and reaction moment M in the directions indicated by the arrows. The reaction force V is directed oppositely to the beam&#39;s primary Y R  deflection, and the reaction moment M tends to twist the ends of the beam about a Z-axis perpendicular to the paper to oppose the bending deflection. The forces and moments applied by the beam into the supports vary depending on amplitude and frequency of the beam&#39;s vibration and on the beam&#39;s size. Because the forces and moments applied by the beam tend to shake the mounts to which they are secured, some of the beam&#39;s vibrating energy is lost. The isolation mass structure in FIG. 1 attempts to isolate the vibrating beam from the end mounts in an effort to attenuate the reaction forces and reaction moments.  
           [0009]    The vibration mode of the vibrating beam in FIG. 3 is that of a virtual-fixed vibrating beam. The term “virtual” is used in the sense that at the beam roots there is a slight Y-axis displacement (Y R ) and a slight angular displacement about the Z-axis as indicated by angle A. Y R  is on the order of one percent of the positive Y-axis displacement from the neutral center line Y R  of the maximum beam in deflection. The angle A can vary but is generally on the order of ±1% of the point of maximum vibrating beam slope.  
           [0010]    The bending moment M and shear force V reactions at the beam root are the primary reactions of the system. Because the spring members  22 - 28  are long and thin, and therefore flexible for Y displacement, the F 1Y  reactions are very small but not completely negligible. However, because these beams are axially stiff, the F 1X  reactions, while small compared to the shear force v and the bending moment M reactions, are not negligible and are a primary cause of energy loss through the end mounts.  
           [0011]    A net F 1X  may result from the F 1X  forces of the two isolator members  18 ,  20  being opposite but unequal. In addition, a net moment reaction may result from a moment reaction due to the F 1X  forces acting in different directions. These net forces or moment reactions are transferred through the isolator members  18 ,  20  to the mounts structure resulting in energy loss and reduced Q. However, both the net F 1X  and the moment effects may be reduced if the axial Y R  displacement and the angular A displacement can be reduced.  
           [0012]    [0012]FIG. 4 is a simplified version of FIG. 3 which considers just the major moment M and shear force V reactions transferred to the isolator mass structure center of gravity (CG) using the well-known parallel axis theorem. Even though the FIX reactions are present and are the major cause of energy loss, they are small compared to the M and V reactions, and therefore, are omitted from FIG. 4 to simplify the drawing and analysis. For this simplified situation, the isolator member reaction to the moment M and shear force V reactions is purely inertial, and the Y R  and A displacements can be closely approximated by equations (1) and (2) set forth below:  
               Y   R     ≈         -   V           m   I          (     2      π                 f     )       2                     and             (   1   )               A   ≈       (     M   -   Vd     )         m   l              k   2          (     2      π                 f     )       2                 (   2   )                               
 
           [0013]    where Y R  is the vertical displacement of the beam root and therefore the isolator mass CG from the neutral center line shown in FIG. 3, V is the shear force, m 1  is the mass of the isolation system (corresponding to the two isolator mass structures), ƒ is the vibration frequency of the beam, M is the bending moment, d is the distance between the isolator mass structure center of gravity and the vertical support member of the isolator mass structure, and k is the radius of gyration of the isolator mass structure. Approximation signs are used instead of equal signs because of the simplifying assumptions used as described above.  
           [0014]    The mass of the isolator structure m 1  appears in the denominator in both equations (1) and (2), which indicates that both Y R  and A are inversely proportional to the mass of the isolator structure. As a result, the greater the mass of the isolator mass structure, the smaller the displacements Y R  and A which results in greater effectiveness of the isolation mechanism and a higher Q for the vibrating beam force transducer.  
           [0015]    Furthermore, equation (2) indicates that the angular displacement A can be reduced to zero if M equals Vd, hereafter referred to as the “tuned condition.” In the tuned condition, the angle displacement A, and hence the F 1X  reactions, approach zero. For a fixed vibrating beam, flexure theory predicts a fixed relationship between M and v such that the condition M=Vd can be brought about if d=0.215 L B , where L B  is the length of the vibrating beam. Proper proportioning of the isolator members should be used to locate the center of gravity of each isolator mass structure to a tuned condition position that cancels the moment M. Unfortunately, the vibration isolator design shown in FIG. 1 does not permit a massive isolator mass that can operate at an optimal tuned condition.  
           [0016]    It is an object of the present invention to provide an effective vibration isolator design that minimizes the amount of energy transferred from a vibrating beam to a mount structure.  
           [0017]    It is an object of the present invention to provide an effective vibration isolator design that minimizes at the vibrating beam roots Y-axis displacement (Y R ) and angular displacement (angle A) about the Z-axis.  
           [0018]    It is an object of the present invention to provide a vibration isolator design with a massive isolator mass structure shaped and proportioned so that a tuned condition operation may also be achieved.  
           [0019]    It is an object of the present invention to provide a vibrating beam force sensor apparatus in which isolator mass structures are skew-symmetric with respect to the horizontal vibrating beam to permit more massive isolator structures with centers of gravity located to achieve an optimal tuning condition.  
           [0020]    The force sensor apparatus includes a vibrating beam and first and second isolator mass members that support the ends of the vibrating beam. The first and second isolator mass members are configured symmetrically relative to an axis that intersects the vibrating beam at an angle other than 90 degrees. First and second end mounts connect respectively to the first and second isolator members. Each isolator mass member has a center of gravity. Each isolator mass member is shaped so that it can be massive (e.g., along the x-axis direction) while at the same time having its center of gravity at an optimal location so that undesirable beam forces and moments that would otherwise transfer vibrating beam energy to the end mounts are cancelled.  
           [0021]    In a preferred example embodiment, each isolator mass member is L-shaped. One example L-shape includes first and second perpendicular portions with the second arm portion extending perpendicular to the first portion and parallel to the vibrating beam. The first and second portions of each isolator mass member are sized such that a distance between the center of gravity and the closest edge of the first portion corresponds to a condition where little or no energy from the vibrating beam is transferred to the corresponding end mount. In a preferred example embodiment, that distance is approximately 0.215 multiplied times the distance between the closest edges of the first portions of the two isolator mass members. A thickness of the first portion and a length of the second portion of each isolator mass may be increased to increase the mass of the isolator while still maintaining the center of gravity for each isolator mass member at the location for optimal “tuned” operation.  
           [0022]    In another example embodiment, another example L-shape includes a third arm portion, considerably shorter than the second portion, that also extends perpendicular to the first portion and parallel to the vibrating beam. Because the third portion is relatively short compared to the second portion, the optimal location of the center of gravity and the increased massiveness of the isolator mass are preserved. Other shapes may be employed as long as there is “skewed” symmetry that allows increased isolator member massiveness while preserving its center of gravity at an optimum, tuned location.  
           [0023]    In the context of a force sensor, electrodes are provided to stimulate the vibrating beam into vibration and for monitoring the frequency of vibration related to a direction and an amount of force applied to the force sensor apparatus. Such a force sensor finds example application as a pressure sensor, an accelerometer, as part of a scale, or any other force sensing environment. 
       
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS  
       [0024]    The foregoing and other objects, features, and advantages of the present invention may be more readily understood with reference to the following description taken in conjunction with the accompanying drawings where like reference numbers refer to like elements.  
         [0025]    [0025]FIG. 1 illustrates a beam force transducer with integral mounting isolation;  
         [0026]    [0026]FIG. 2 illustrates the force-frequency effect of a vibrating flexor beam;  
         [0027]    [0027]FIG. 3 illustrates an exploded view of the beam structure of FIG. 1 in flexure;  
         [0028]    [0028]FIG. 4 illustrates a simplified drawing of the isolator mass structure;  
         [0029]    [0029]FIG. 5 illustrates an isolator mass structure in accordance with one example embodiment of the invention;  
         [0030]    [0030]FIG. 6 illustrates an isolator mass structure in accordance with another example embodiment of the invention; and  
         [0031]    [0031]FIG. 7A is a top view while FIG. 7B is a side view of a pressure sensor assembly incorporating the isolator mass structure of FIG. 5. 
     
    
     DETAILED DESCRIPTION  
       [0032]    In the following description, for purposes of explanation and not limitation, specific details are set forth, such as particular embodiments, techniques, etc. in order to provide a thorough understanding of the present invention. However, it will be apparent to one skilled in the art that the present invention may be practiced in other embodiments that depart from these specific details. In some instances, detailed descriptions of well-known methods, devices, and techniques are omitted so as not to obscure the description of the present invention with unnecessary detail.  
         [0033]    The present invention may be used in any vibratory beam apparatus. While the vibrating beam apparatus may be formed from any single material, it is preferred that the apparatus be made of a single block of quartz crystal or other piezoelectric material. While the preferred material is quartz crystal, other metallic or non-metallic material can be used. If these materials are not piezoelectric, an alternate vibrating drive mechanism would be used. Example applications of the present invention are in the context of accelerometers, pressure sensors, scales, etc. The invention may be used in other force sensing applications as well.  
         [0034]    [0034]FIG. 5 illustrates a vibrating beam apparatus  50  in accordance with an example embodiment of the present invention. The vibrating beam apparatus  50  includes some structures similar to those described in FIG. 1 including end mounts  12  and  14 , a vibrating beam  16 , isolator members  18  and  20 , coupled to respective ones of the end mounts by corresponding thin spring members  22 ,  24 ,  26 , and  28 . The electrodes on the surface of the vibrating beam  16  and wires to an oscillator drive circuitry are not shown to simplify the drawing. However, the isolator members are indicated as  18 ′,  20 ′ because they are shaped and sized differently from the members  18  and  20  shown in FIG. 1.  
         [0035]    In the isolator mass design shown in FIG. 1, the isolator members are symmetric relative to plural axes parallel to the y-axis and perpendicular to the y-axis. In FIG. 5, the isolator members  18 ′,  20 ′ are symmetric in a “skewed” fashion relative to an axis that intersects the vibrating beam (parallel to the x-axis) at an angle other than 90 degrees. The term “skewed” means slanted or the like, and one can view the isolator members  18 ′,  20 ′ as “skew-symmetric” or symmetric relative to a slanted line.  
         [0036]    In the example embodiment of FIG. 5, each isolator member  18 ′,  20 ′ includes an extended isolator mass arm. Isolator member  18 ′ includes a vertical base portion  54  and a horizontal arm portion  55  in a first configuration. Isolation member  20 ′ includes a vertical base portion  52  and a horizontal arm portion  53  in a second configuration. The first and second configurations of the isolation members  18 ′,  20 ′ are symmetric with respect to the slanted axis  60  that intersects the beam  16  at an angle other than 90 degrees. The skew-symmetric design allows for the base portions  54  and  52  to be relatively thick in the X direction to increase the mass of each isolator member  18 ′ and  20 ′. At the same time, the skew-symmetric design also allows each of the horizontal arm portions  55  and  53  to extend far over the vibrating beam  16  in opposing parallel planes to counter the increased mass of the thicker base portions  54  and  52 . The extended arm portions  55  and  53  maintain the centers of gravity  58  and  56  of the isolator mass structures for each of the isolation members  18 ′ and  20 ′ at the optimum position for tuned condition operation.  
         [0037]    Using the skew-symmetric design, the mass m 1  of each isolator member in equations (1) and (2) can be increased substantially. From those equations, an increase in the mass of the isolator mass members decreases the axial and angular displacements Y R  and A resulting in greater effectiveness of the isolation mechanism and a higher Q for the vibrating beam force transducer. Because the arm portions overhang the beam  16  to a considerable extent, each isolation member can be designed so that the moment M equals the shear force V times the distance “d” in the X direction between the center of gravity  56  and  58  of the isolator mass structures and the point where the beam  16  connects to each of the isolator members  18 ′,  20 ′. The optimum distance d to achieve the tuned condition, i.e., M=Vd, is when d= 0 . 215 L B , where L B  is the length of the vibrating beam  16 . Operating at the tuned condition reduces the angular displacement A to theoretically zero. Making the vertical base portions thicker in the X direction moves the center of gravity for each of the isolator members closer to the base portion reducing d to a value less than the optimum 0.215 L B . However, by extending the horizontal arms  55  and  53  in the X direction, the center of gravity for the isolator mass structures for each of the isolator members  18 ′ and  20 ′ is moved in the opposite direction, thereby maintaining the distance d at the tuned condition of 0.215 L B .  
         [0038]    [0038]FIG. 6 illustrates another example skew-symmetric design for a vibrating beam apparatus  50  that is similar in most respects to the design shown in FIG. 5. However, the isolator mass members  18 ′,  20 ′ each include a third arm portion  62  and  64 , respectively, that extend perpendicularly from their respective base portions  54  and  52  for a short distance in a plane parallel to the vibrating beam  16  and to their respective second arm portions  55  and  53 . The isolator mass members  18 ′,  20 ′ are symmetric about a skewed line  60 ′ and permit, in similar manner to the design in FIG. 5, increased isolator member massiveness while still preserving the optimum, tuned condition location of the centers of gravity  58  and  56 . Other skew-symmetric designs may be employed with similar benefits.  
         [0039]    [0039]FIGS. 7A and 7B illustrate an example application of the vibrating beam force sensor  50  shown in FIG. 5 in the context of a pressure sensor located generally at  100 . FIG. 7A includes a top view and FIG. 7B a side view of the pressure sensor  100 . A skew-symmetric vibrating beam force sensor  50  is coupled to and is a part of a crystal resonator  110  secured by a mount screw  114  to a sensor housing  102 . The crystal resonator  110  is enclosed and evacuated in sealed housing. Locating the structure in an evacuated housing avoids air resistance which would otherwise dampen vibrations and reduce the Q of the vibrating beam of sensor  50 . The crystal cavity is sealed using a top cover  110  and a bottom cover  122  by braising, welding, soldering, or the like. A getter  116  is included in the crystal cavity to maintain vacuum quality. Evacuation is achieved by way of an exhaust tube  118 . Electrical feed-throughs  112  are provided for wires to connect beam electrodes plated on the resonator to an oscillator circuit board  124  shown in FIG. 7B. Pulses from the oscillator  124  to the electrodes cause the beam to vibrate at a particular frequency.  
         [0040]    A bellows  108  is made of electrode-deposited nickel with a thin wall thickness for very low spring rate. The conical termination of the bellows  108  forms a well-defined contact point where it meets one of the end mounts  134  of the crystal resonator  110 . The bellows  108  is coupled at its other end to a fitting  104  inserted into the housing  102  via a hub  106 . Orthogonal flexure beams  132  permit the end mount  134  and a balance weight  136  to rotate about pivot point  130  under the influence of pressure to the bellows  108 . The bellows  108  converts fluid pressure to a force acting upon the end mount  134 . The force is caused by a pressure difference between fluid inside and outside of the bellows  106 . Movement about the pivot point  130  is resisted by the vibrating beam experiencing a compression force which changes the resonant vibrating frequency of the vibrating beam. The change of frequency is thereby a measure of the fluid pressure.  
         [0041]    While the present invention has been described with respect to particular embodiments, those skilled in the art will recognize that the present invention is not limited to these specific exemplary embodiments. Different embodiments and adaptations besides those shown and described as well as many variations, modifications, and equivalent arrangements may also be used to implement the invention. Therefore, while the present invention has been described in relation to its preferred embodiments, it is to be understood that this disclosure is only illustrative and exemplary of the present invention. Accordingly, it is intended that the invention be limited only by the scope of the claims appended hereto.