Patent Publication Number: US-2019186575-A1

Title: Planar Linkage, Methods of Decoupling, Mitigating Shock and Resonance, and Controlling Agricultural Spray Booms Mounted on Ground Vehicles

Description:
This United States utility patent application is a continuation application of pending application Ser. No. 15/299,383 filed Oct. 20, 2016, which itself is a divisional application of application Ser. No. 14/213,145 filed Mar. 14, 2014 (now U.S. Pat. No. 9,504,211 issued Nov. 29, 2016), which itself claims priority on and the benefit of expired provisional application 61/794,655 filed Mar. 15, 2013, the entire contents of all are hereby incorporated herein by reference. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to a planar linkage, to the support, isolation and dynamic control of spray booms connected to a tractor, trailer or other vehicle for the purpose of agricultural spraying and to a gas tension spring. 
     BACKGROUND OF THE INVENTION 
     Agricultural fields are often sprayed with various spraying solutions, such as herbicides, insecticides, fertilizers, etc. Sprayers for this purpose are required to have a wide span booms furnished with plumbing and multiple spray nozzles, for the liquids being sprayed, at defined intervals across the boom&#39;s span: While contemporary span widths may extend as wide as  120  through to 162 feet, with even wider spans predicted in the future. However, when not being used for spraying, these wide-span booms are required to fold away to a stowed position, typically along either side of the vehicle, to permit the vehicle to navigate field entrance gates and traverse tracks, roads or highways, without exceeding either practical or legal width limits. 
     The vehicles which employ these foldable wide span booms may take the form of farm tractors, trailers or specialist vehicles fitted with chemical tanks or reservoirs to carry the liquids being sprayed, and the spray booms themselves may be fitted typically at either the front or rear of such vehicles on each side of the vehicle on a “center-rack” which also supports plumbing and spray nozzles across the short distance that spans the width of the vehicle itself. Typically, but not universally, this center-rack, along with the booms attached to it, can be elevated or lowered on a four-bar linkage and actuator(s) or other suitable means, to adjust and set the boom/nozzles spray height to achieve to most effective crop spray coverage. Typically, at their attachment to the outer edges of the center-rack, the booms are arranged to pivot about essentially vertical axes, so that they may be rotated in angular displacement through 90° or so from the spraying position which is normal to the vehicle&#39;s longitudinal axis, to the folded, stowed position essentially parallel the vehicle&#39;s longitudinal axis when viewed in plan. Because of the great length of the booms, it is customary for the booms to be further folded via hinge points located at around the mid-semi-span of each boom. In some cases, where very high span booms are used, further folding hinge points may be used to shorten the folded boom length. In other cases, the folded boom length may be shortened by having the outer semi span of each boom retract telescopically in to the inner boom semi-span. In yet others, a combination of telescopic and folding boom segments may be employed. In most cases, the folding and unfolding action of the booms is conducted by means of actuators, more commonly, hydraulically operated. Conveniently, when the full span of the booms may not be required (because of width limitations of parts of the field being sprayed, for example) spray booms are commonly designed to be operated in the part-folded position, with the outboard sections folded alongside the inboard sections. Thus a boom system might be referred to as a  132 - 60 , or other similar designation, implying in this case a full span of 132 feet, and a semi-folded span of 60 feet. 
     The loads imposed on the booms during operation have a significant effect on their structural design. It is the mass of the booms&#39; structure itself along with its supported loads including pipework, plumbing, spray nozzles, valves, filters, hydraulic cylinders and sundry masses such as touch-down wheels that is responsible for generating the greater part of its structural loading. 
     Since the mass of the boom assemblies, as described above, acted on by gravity and by inertial accelerations in the vertical direction, about the roll axis and by inertial and some gravitational components in the longitudinal direction as well as about the yaw axis; maximizing the specific strength of the booms by minimizing their mass relative to their structural strength is an extremely important, if not critical aspect of high-span spray boom design. It is therefore highly advantageous to design high-span spray booms using high-strength lightweight materials and to incorporate specific design features that simplify and aid manufacture, while keeping costs to a minimum. 
     Of the destructive loads able to be imposed upon the deployed booms by the movement of the vehicle as it traverses the undulating surface of the farm land being sprayed, the vehicle&#39;s movement about the roll and yaw axes are potentially the greatest. This is because the booms effective moment of inertia about these axes can be defined as the sum of an infinite number of discrete mass segments each of whose moment is the product of the segment mass and the square of its distance from the roll axis: Therefore because of the large boom span and the distance-squared function, the polar moment of inertia of the deployed booms is truly massive, notwithstanding that the booms&#39; outboard sections can be comparatively light. Accordingly, if the vehicle moves in angular displacement about its roll axis due to continuously varying relative vertical displacements of the wheels at either wheel track, then enormous potentially destructive forces may be generated at the booms&#39; attachments to the center rack, unless some mitigating design features are incorporated to prevent or reduce such forces. 
     One way that this is currently done is to arrange to allow angular displacements to readily occur between each inner boom and the center-rack at a lower, acceptable level of force, by having a lower longitudinal pivot between the boom and center-rack. Each boom is then maintained in an essentially horizontal position by a hydraulic cylinder that attaches to an upper inboard boom attachment point at one end and to an upper center rack attachment at the other end. By incorporating relatively small pressurized gas accumulators in the positively pressurized hydraulic supply lines to these cylinders, a level of compliance (springing) in angular roll displacement can be achieved between each of the booms and the center-rack/vehicle. Further, by arranging for a control system to lengthen or shorten the each upper boom attachment cylinder, each boom can be controlled independently in angular displacement about the vehicle&#39;s roll axis; and this can be used, in conjunction with ground-height sensors mounted, typically, at intervals across the span of the booms, as part of a system to control and maintain to booms essentially parallel to the ground during operation. It also permits control recognition of sudden vehicular roll movements and allows active correction of the boom position relative to the vehicle&#39;s roll position when such spurious movements occur. While this has proven to be effective, at least to some extent, the active control response is often considered to be too slow to be fully effective in mitigating the movements and forces caused by sudden roll excursions of the vehicle. Consequently, over the more undulating surfaces, the vehicle speed may have to be reduced to unacceptably slow levels to allow time for the corrective response to take place, or loss of effective control of the outboard boom section heights may take place, giving rise to unnecessarily high boom forces in roll as well as defective spray application and may even cause the boom tips to impact the ground. 
     Again, according to contemporary practice, the foregoing active roll correction system is sometimes further improved by linking the hydraulic lines feeding two upper boom hydraulic cylinders on either side of the vehicle together via pressure relief valves: In the event that the roll forces imposed on the cylinders gives rise to a hydraulic pressure differential between the cylinders that exceeds a given pre-set level, the relief valves then open and hydraulic fluid is transferred automatically between the two cylinders, allowing the vehicle to effectively roll relative to the pair of booms en-masse without having to react further forces. This may well be an improvement, but it is not a full solution since typically outer boom height control is still rendered somewhat ineffective in practice. 
     An alternative way that the potentially destructive loads caused by the vehicle&#39;s roll excursions from reacting the polar moment of the booms is currently addressed, is to mount the center-rack, to which the booms are attached, on a longitudinally aligned pivot at the center-rack&#39;s primary attachment to the vehicle. A torsionally resilient connection may be used at this point and this may take the form of torsionally acting spring elements or other means to help keep the boom in general horizontal alignment, relative to the vehicle, without the vehicle&#39;s short term roll movements significantly deflecting the booms in roll, or generating excessively high reaction forces in the booms at their attachment to the center-rack. A further variation on this theme is for the center-rack to be supported by a linkage that results in an effective virtual longitudinal pivot point whose virtual pivotal axis is above the center of gravity of the combined booms and center rack, such that the booms benefit by the pendular stability so generated, at least when travelling on fairly level terrain. The upper boom attachment hydraulic cylinders, or a single hydraulic cylinder and linkage serving to replace them, then acts to change the angular displacement of the pair of booms in roll relative to each other, rather than relative to the vehicle. 
     A second, independent, control action may then be employed to control the overall position of the linked pair of booms in roll, relative to the ground reference, given by the previously mentioned boom height sensors. Thus, by controlling these two sets of boom roll position criteria, the booms may not only have the roll forces, otherwise imposed on them by the vehicle movement over undulating or rough ground surfaces, reduced or effectively eliminated, but a comprehensive boom height control system can effectively permit the booms to be maintained at an essentially fixed mean-height above the ground; and also that this mean height can be maintained even when traversing the rounded crest of a hill or ridge, or along a gully by virtue of being able to control the roll position of the booms relative to each other at the same time. Thus, the spray booms are better able to follow, at an essentially constant mean height above the ground, any gently varying contours of the ground that occurs across the span of the booms during operation. 
     Again, according to contemporary practice, there are two recognized methods by which the active control force can be applied to control the mean angular position of the linked pair of booms relative to the ground: one is to react the controlling actuator in roll against the vehicle, while typically incorporating an interposed low spring-rate compliant element, such that the reaction force is rendered at least somewhat independent of the relative angular roll axis position of the vehicle: While the other is to change the lateral position of the combined booms&#39; center of gravity relative to the center-rack&#39;s roll pivot support, such that gravitational reaction is used. This latter concept has the advantage of deriving the boom roll control forces entirely independently of the vehicle&#39;s instantaneous roll position. This can be achieved, for example, by displacing weights slidably attached to the booms, laterally in order to apply corrective roll forces. One example of such an arrangement is disclosed in WO2012146255, “Active Damping System for a Spray Boom”, Maagaard Jorgen, 2012. 
     On a practical note, one boom design feature that has become almost universally adopted by current wide-span spray boom designs is the “breakaway”. This is typically a vertical hinge pivot system applied such that the last outboard 12 to 15 feet or so of the boom, up to the boom tip itself, can pivot back to alleviate damage if the outermost extremities of the boom accidently contact an obstacle, or contact the ground. There are a number of ways that this is achieved in practice, one most common one being of the double pivot “saloon-door” hinge type, where the breakaway section is centered in the fully extended position by pin inclination and gravity or by spring force, or both, so that upon contact with an object, the breakaway section fold back to avoid damage, and re-centers automatically when the object or ground contact has passed. 
     Another practical adaption often used on wide-span agricultural spay booms are so called “touch-down” wheels. These wheels are attached on legs, one on each boom semi-span below and slightly forward of the booms to avoid interference with the spray pattern, fairly well outboard along the boom span. Their purpose is to prevent the booms from encroaching too close to, or touching the ground in the event of the control system failing to adequately maintain the correct height position of the boom. While such touch-down wheels may prevent obvious damage to the booms in the event the height control system failure, their inclusion might be considered as indication of the inadequacies of current spray boom/control system design and control methodologies and the need to address them. 
     Structural design is of vital importance to both the affordability and durability of wide-span spray booms. In this respect it is not only the absolute structural strength of the booms that is relevant, but also, and perhaps more critically, the fatigue strength, which on metal boom structures, particularly welded metal boom structures, usually defines the boom&#39;s usable life. In this respect the amplitude of the cyclic fatigue loadings applied to the boom, either as imposed loads (from bumps in the terrain reacting the inertia of the boom structure, for example) or as resonance generated loads (from structural modal resonance response) are of great importance. This is because the characteristic fatigue S-N curves (cyclic Strain amplitude verses Number of strain reversals to failure) follows a logarithmic curve with a slope of approximately three, so effectively represents a number of cycles to failure that varies inversely as the cube of the cyclic strain. To put this into perspective, if by the severity of operation, the magnitude cyclic loading forces on a given boom structure were to be doubled, then its fatigue life would be expected to fail prematurely at around just one eighth of its original value. While, on the other hand, if by design, the cyclic loading were to be halved, then the same boom would be expect to benefit by an eightfold increase its life. 
     The alleviation of fatigue loads by adequate compliant suspension in heave (vertical accelerations imposed when traversing undulating or bumpy ground), in roll (which has been addressed in the foregoing paragraphs), in longitudinal acceleration (acting inertially to flex the booms backwards and forwards on accelerating and braking or climbing or descending gradients) and in yaw (accelerations imposed about the yaw axis by steering the vehicle), is commonly practiced in current designs. In some cases, semi-active control of the longitudinal and yaw accelerations is also currently practiced, while automatically self-leveling the booms relative to the sensed ground position at a pre-set spray height, combined with compliant boom suspension, effectively results in semi-active vertical boom suspension, there still remain some serious deficiencies in structural and fatigue boom design capabilities. 
     Primarily these relate to the propensity to structural resonance in the (necessarily very flexible) booms excited by vertical and/or longitudinal accelerations in the supporting vehicle due to its operation over rough or undulating ground, even when the best methods of trying to isolate the booms from such critical vibration frequencies have been employed. Such resonant vibrations, magnified by an exciting frequency, can rapidly fail or fatigue the boom&#39;s structure prematurely. Further, designing to avoid the critical frequencies generated by the vehicle is largely thwarted by the potentially wide range of frequencies able to be generated due to mass of the vehicle changing on its suspension and tires, as its liquid cargo is discharged during the spraying operation. This is a significant weakness in contemporary high-spans pray booms, and one which will only become worse as economic necessity drives future spray boom spans wider. 
     The optimal structural design of spray booms typically results in a triangulated braced truss-structure for several reasons. Firstly, the truss type structure is one of the strongest, lightest and most rigid configurations, and secondly, when in the folded position along both sides of the vehicle, the open lattice frame of the truss structure allows the driver a fairly high level of visibility through the structure itself, so enabling safer operation, particularly on roads and highways. The open lattice structure also permits ready access to any plumbing, hydraulics, electrical and communication lines, sensors etc. for maintenance or modular adaptability. 
     From the foregoing it can be seen that there are a large number of relevant factors that need to be addressed in the optimal design of wide-span spray agricultural booms, and that contemporary designs are deficient in a number of respects. 
     It is desirable that wide-span spray booms be designed using lightweight high-strength materials, so that the booms&#39; span can be maximized while the structural loads, resulting largely from the boom structural mass, can be kept with the limits defined by the operational life requirements. 
     It is desirable that, in order to maximize agricultural sprayer utility in terms of area sprayed in unit time that both the boom-span and vehicle speed be maximized; notwithstanding that both these parameters significantly increase the propensity of the boom structure to flexure and resonance. 
     It is desirable for the boom system to be able to accurately maintain the optimal, near constant, spray height above the ground, and to follow the smooth contours and undulations in the ground surface profile in span at the highest practical vehicle spraying speed. 
     It is desirable that not merely the limit-load strength of the boom structure, but the fatigue strength of the structure, be primary criteria for spray boom design. 
     It is desirable that the problem of boom structural resonance, particularly in the vertical and longitudinal vehicle axis directions, be eliminated or reduced to acceptable levels, particularly at the resonant Eigen-frequencies. 
     It is desirable that the spray boom structure be of the truss or lattice type, so that the vehicle driver&#39;s visibility through the structure is not significantly impeded when the booms are in the folded position along both sides of the vehicle. 
     It is desirable that methodologies to mitigate otherwise excessive loads from being imposed on the booms&#39; structure and attachments to the vehicle due to the vehicle&#39;s angular displacements in roll over uneven ground being reacted against the deployed booms&#39; extremely high polar moments of inertia in roll 
     It is desirable that methodologies to mitigate otherwise excessive loads from being imposed on the booms&#39; structure by the vehicle&#39;s movement over rough or uneven ground, in the vertical or longitudinal axis directions due to the booms&#39; inertia reacting the vehicle&#39;s vertical, longitudinal and yaw displacements. 
     It is desirable that, in the design of the booms combined with their attachments to their supporting center-rack, along with the center-rack&#39;s attachment to the vehicle, that provisions be made to support the use of advanced active boom control methodologies: These are methodologies that enable the booms to follow the varying contours of the ground with a high level of accuracy, without interference from spurious short term vehicle displacements, and with a response time consistent with these objectives. 
     The present invention serves to overcome these deficiencies. 
     SUMMARY OF THE INVENTION 
     The current invention discloses several methodologies that mitigate shock loading and propensity to resonance in agricultural spray boom structures. These include a new form of near-planar linkage instrumental in decoupling the boom assembly from the vehicle in pitch, heave and roll. This serves to permit further aspects of the invention to: Use the combined mass of the booms and center section as the tuned mass of a tuned mass damper that can counter the eigenfrequency of the boom system in vertical resonance (flapping): Act as an enabling part of a boom compliant suspension system to mitigate shock loadings otherwise imposed on the boom system, and: Act as an enabling part of an active boom height and roll control systems to permit the accurate navigation of the boom over undulating terrain. The planar linkage also has application to other devices and uses. Further aspects of the current invention include the incorporation of tuned mass dampers to counter resonance at either the Eigen or tertiary frequencies in the boom structure and components; and the use of the mass and operation of the boom outboard “breakaway” sections as tuned mass dampers to counter modal resonance in the horizontal plane. 
     There are many advantages of the present invention. Some advantages are: 
     According to one advantage of the present invention, wide-span spray booms of the present invention are designed using lightweight high-strength materials. This allows the boom&#39;s span to be maximized while the structural loads, resulting largely from the boom structural mass, can be kept with the limits defined by the operational life requirements. In order to maximize agricultural sprayer utility in terms of area sprayed in unit time that both the boom-span and vehicle speed be maximized; notwithstanding that both these parameters significantly increase the propensity of the boom structure to flexure and resonance. The fatigue strength of the structure in addition to the limit-load strength of the boom structure should be accounted for in a boom design. Accordingly, an advantage of the present invention is that structural resonance, particularly in the vertical and longitudinal vehicle axis directions, can be eliminated or reduced to acceptable levels, particularly at the resonant Eigen-frequencies. 
     The resulting fatigue strength and structural strength allow for the booms to have a greater width and the vehicle to travel at a faster speed. This results in greater coverage per unit time. 
     According to another advantage of the present invention, the spray boom structure is of the truss or lattice type, so that the vehicle driver&#39;s visibility through the structure is not significantly impeded when the booms are in the folded position along both sides of the vehicle. 
     A near planar linkage is provided according to one aspect of the present invention. The near planar linkage can carry weights in a plane without the need for direct contact points such as rollers. It is understood that the motion of an object connected to the linkage can be nearly planar. 
     According to another advantage of the present invention, the boom system is able to accurately maintain the optimal, near constant, spray height above the ground, and can follow the smooth contours and undulations in the ground surface profile in span at the highest practical vehicle spraying speed. 
     According to another advantage of the present invention, methodologies to mitigate otherwise excessive loads from being imposed on the boom&#39;s structure and attachments to the vehicle due to the vehicle&#39;s angular displacements in roll over uneven ground being reacted against the deployed booms&#39; extremely high polar moments of inertia in roll are provided. 
     According to another advantage of the present invention, methodologies to mitigate otherwise excessive loads from being imposed on the boom&#39;s structure by the vehicle&#39;s movement over rough or uneven ground, in the vertical or longitudinal axis directions due to the booms&#39; inertia reacting the vehicle&#39;s vertical, longitudinal and yaw displacements are provided. 
     According to a further advantage of the present invention, that in the design of the booms combined with their attachments to their supporting center-rack, along with the center-rack&#39;s attachment to the vehicle, that provisions are made to support the use of advanced active boom control methodologies: These are methodologies that enable the booms to follow the varying contours of the ground with a high level of accuracy, without interference from spurious short term vehicle displacements, and with a response time consistent with these objectives. 
     According to an advantage of the present invention, when a center rack forms a tuned mass damper, the entire boom is damped without adding appreciable weight to the system. 
     According to another advantage of the present invention, there is a relatively large reservoir coupled with a bag having a small volume. This allows for the bag to provide a near constant force to support the boom. 
     Further, the near constant force element is coupled with dampers in parallel to eliminate low rate of oscillation in booms. Hence, the entire boom is damped. 
     According to a further advantage of the present invention, the angle of the center rack (and hence the booms) can be controlled by an internal pump. A controller controls an internal pump to direct air (and hence change pressure) to rotate the boom assembly to the left or right side boom. This is accomplished without adding or removing air (or a gas or a working fluid) from the system. The internal pump switches direction of pumping as often as needed to maintain the desired boom angles. 
     According to another advantage of the present invention, a positional link is provided. The positional link can tilt a secondary section of boom relative to the first section of the boom to maintain a desired spray height. Articulation can be done in real time by the use of sensors and a controller. 
     According to another advantage of the present invention, a tension gas spring is provided. The tension gas spring inverts action a compression gas strut to form a tension gas spring. This allows an embodiment of the present invention to pull two objects towards each other. 
     Damping is provided as two arms pull back towards each other in latter portion of return of the gas strut. It is appreciated that there isn&#39;t any appreciable damping as the arms separate from each other, thereby allowing freedom of outward motion. 
     In use, two tension gas springs can be used with a breakaway section of a boom to turn the break away into a tuned mass damper. In this regard, depending on the direction of the swing, one of the two tension gas springs can first separate (without damping) and then provide a damping effect upon the latter part of the return. Hence, there will be one active damper and one inactive damper depending on the direction of the swing of the breakaway relative the adjacent boom section. 
     Another advantage of using the breakaway section as a tuned mass damper is that it does not add appreciable weight to the boom. 
     According to another aspect of the present invention, a tuned mass damper can be bolted or otherwise connected to the boom. This advantageously can address specific problems in primary and secondary (or tertiary) modes remedially (i.e. if they appear). 
     According to another advantage of this aspect of the present invention, the tuned mass damper can be designed so that a single mass can damp in both the vertical and horizontal directions. 
     According to another advantage of the present invention, the tuned mass dampers can be passive or active. In an active damper, such as a coil and magnetized mass, the user can selectably turn on the damper as necessary. 
     Other advantages will become apparent by studying the following detailed descriptions and the drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a perspective view of an existing agricultural spray vehicle and boom system in operating position to which the current invention may be applied. 
         FIG. 2  is a perspective view of an existing agricultural spray vehicle in folded and boom system, to which the current invention may be applied, with the booms in the stowed position. 
         FIG. 3  is a perspective view of an embodiment of a near planar linkage. 
         FIG. 4  is a side view of the embodiment illustrated in  FIG. 3 . 
         FIG. 5  is similar to  FIG. 4 , but shows a strut moved away from the centered position but maintaining a position in a similar plane. 
         FIG. 6  is a rear perspective view of the embodiment illustrated in  FIG. 3 . 
         FIG. 7  is similar to  FIG. 6 , but shows a spherical joint at the distal end of the strut in an alternative orientation. 
         FIG. 8  is a perspective view of an embodiment of a support assembly including a center rack incorporating near planar linkages. 
         FIG. 9  is an alternative view of the embodiment shown in  FIG. 8 . 
         FIG. 10  is an alternative view of the embodiment shown in  FIG. 8 . 
         FIG. 11  is an alternative view of the embodiment shown in  FIG. 8 . 
         FIG. 12  is a perspective view showing an embodiment of a positional connector of the present invention. 
         FIG. 13  is similar to  FIG. 12 , but shows the second boom section in an elevated angle relative the primary boom section. 
         FIG. 14  is similar to  FIG. 12 , but shows the second boom section in a declined angle relative the primary boom section. 
         FIG. 15  is a high level flow diagram of the controller operation. 
         FIG. 16  is a side view of an embodiment of a tension gas spring. 
         FIG. 17  is a perspective view of the embodiment illustrated in  FIG. 16 . 
         FIG. 18  is a side view showing the internal components of the embodiment illustrated in  FIG. 16 . 
         FIG. 19  is a perspective view of a breakaway tuned mass damper. 
         FIG. 20  is similar to  FIG. 19 , but shows the breakaway swung in a first direction. 
         FIG. 21  is similar to  FIG. 19 , but shows the breakaway swung in a second direction. 
         FIG. 22  is a perspective view of an embodiment of a tuned mass damper. 
         FIG. 23  is a perspective view of an alternative embodiment of a tuned mass damper. 
         FIG. 24  is a perspective view of an alternative embodiment of a tuned mass damper. 
         FIG. 24A  is similar to  FIG. 24  but shows a cover in place. 
         FIG. 25  is a perspective view of an alternative embodiment of a tuned mass damper. 
         FIG. 25A  is similar to  FIG. 25  but shows a cover in place. 
         FIG. 26  is a schematic view of a boom assembly with a left and right boom in a straight position. 
         FIG. 27  is similar to  FIG. 26 , but shows the booms articulated to match the contour of the ground beneath. 
         FIG. 28  is similar to  FIG. 26 , but shows the booms articulated to match the contour of the ground beneath. 
     
    
    
     DETAILED DESCRIPTION OF THE INVENTION 
     While the invention will be described in connection with one or more preferred embodiments, it will be understood that it is not intended to limit the invention to those embodiments. On the contrary, it is intended to cover all alternatives, modifications and equivalents as may be included within the spirit and scope of the invention as defined by the appended claims. 
     Referring now to the invention in more detail,  FIGS. 1 and 2  show the booms mounted in position on a vehicle.  FIG. 1  shows them as they would be in the deployed in the full-span and part-span operating position when used for spraying crops, while  FIG. 2  shows the invention in its folded position as it would be used for driving to and from the fields being sprayed, maneuvering through field entrances gates, along tracks or on roads or highways. Unless otherwise noted, the booms  110  and  120  shown in the various embodiments of the current invention comprise three primary identifiable segments: The primary or inner boom, the secondary or outer boom, which incorporates a breakaway. 
     In order to decouple the entire boom system and center-rack from the vehicle in pitch heave and roll, the current invention incorporates multiple near planar linkages, or near planar link mechanism,  20 . These linkages are seen in  FIGS. 3-7 . 
     Each near planar linkage  20  has a body  30  with three body joints  31 ,  32  and  33 . A center arm  40  is provided having a base  41  with three center arm joints  42 ,  43  and  44 . A strut  50  upstands from the base  40 . The strut  50  has a proximal end  51  and a distal end  52 . A spherical joint  53  is at the distal end  52  of the strut. A link  60  is provided. A spherical joint  62  is at a first end  61  of the link, and a spherical joint  64  is at the second end  63  of the link. A link  70  is provided. A spherical joint  72  is at a first end  71  of the link, and a spherical joint  74  is at the second end  73  of the link. A link  80  is provided. A spherical joint  82  is at a first end  81  of the link, and a spherical joint  84  is at the second end  83  of the link. 
     The links  60 ,  70  and  80  interconnect the body  30  and center arm  40 . The first end spherical joint  62  of link  60  is connected to the first body joint  31 . The first end spherical joint  72  of link  70  is connected to the second body joint  32 . The first end spherical joint  82  of link  80  is connected to the second body joint  33 . The second end spherical joint  64  of link  60  is connected to the center arm joint  42 . The second end spherical joint  74  of link  70  is connected to the center arm joint  43 . The second end spherical joint  84  of link  80  is connected to the center arm joint  44 . 
     It is understood that the center arm  40  can more relative the body  30 . The distal end  52  of the strut  50  moves generally in an approximate plane  90 . 
     The central arm in one example can have a length of about 12 inches measured from the center of the spherical joint at the distal end to the plane of the pitch circle diameter (PCD) of the three other spherical joints at its large end. Relative to the central arm length, the link lengths center to the center of the spherical joints can be approximately 0.63; the center rod lard end PCD of the three spherical joints is approximately 0.67; and the large ring PCD of the three spherical joints is approximately 1.17. 
     Stated in more particularity, the near planar linkage comprises three links each of which is attached at one end by a spherical joint to a main body or support  30  at three approximately equi-distant attachment locations, and to one end of a mid-positioned, longer strut, via three approximately equi-spaced spherical joints. By way of definition, it can be stated that the center strut, having the three links attached at one end as shown, and a single spherical joint at its distal end, has a longer distance between the plane of the pitch circle diameter of the spherical attachments at its large end to the center of the spherical joint at its distal end, than the length of the three links between their spherical joint centers: The pitch circle diameter of the spherical attachments on the main support is larger than the pitch circle diameter of the three spherical attachments at the end of the center strut. The length of the three links, spherical joint center to spherical joint center, is less than the pitch circle diameter of the centers of the three spherical joints on the larger end of the center-strut. Configured appropriately, the path that the center of the spherical joint at the distal end of the center-strut follows always remains closely planar to the plane that passes through the pitch circle diameter of the centers of the three spherical joints on main support. However, it is not a perfect planar linkage; there must always be some small deviation: There is so single perfect mathematical solution. The mathematical method by which the geometry of the near-planar linkage is generated therefore one of iterative “Design of Experiments” applied typically using computing processes such as Matlab, Mathcad or bespoke computer programming. For this particular embodiment of the invention, which is not necessarily fully optimized in planar movement and which permits a 12 inch inclusive movement of the center of the spherical joint at the distal spherical joint in all directions, that geometrical deviation is just seventy two thousandths of an inch. This small deviation is considered insignificant in terms of the relative flexibility achievable in the supporting structure of the center-rack to which it is applied in this particular usage. 
     While the foregoing depicts just one embodiment of the near planar linkage used in this particular context, there are numerous potential adaptions that would also be considered to benefit from the use of the concept, A particular aspect is that this specification uses the term “spherical joint” as a way of defining the function of a joint mechanism. Clearly, the effect of a spherical joint function can be simply replicated by the use of multiple revolute joints, for example, a universal, or Hook&#39;s joint; a roller ball joint, CV joint, recirculating ball or roller joint, or indeed, any elastomeric joint configured to achieve the same objective: These are all considered to be synonymous and inclusive with the term “spherical joint”, for the purposes of this specification, as they allow rotation in all directions. 
     These include the application of such a planar linkage to the support of major structures such as high-rise buildings in earthquake prone geographical zones, the general enhancement of tuned mass damping technology, and vehicle suspensions, also expounded in this patent specification, and in many advanced technological areas. 
     Turning now to  FIGS. 8-11 , it is seen that a support assembly  150  is provided to connect to a vehicle  101  four bar linkage  102 . The vehicle has a lift actuator  102 . The vehicle can have a first boom  110  with an angle actuator  111  for controlling an angle of the boom relative to the center rack and a second boom  120  with an angle actuator  121  for controlling an angle of the boom relative to the center rack. 
     The support assembly  150  has a center rack support frame  160 . Support frame  160  has four connectors  161 ,  162 ,  163  and  164 . A center rack  170  is further provided having a top  171 , a bottom  172 , sides  173  and  174 , a front  175  and a rear  176 . A plurality of near planar linkages, preferably four such linkages  180 ,  181 ,  182  and  183 , are provided. The four linkages are removably secures to the connectors of the support frame, wherein the center rack can move within a plane relative the support assembly without appreciably changing the distance between the components, namely the support frame  160  and the center rack  170 . 
     Several other components are provided, including a controller  200 , a feed pump  210 , an internal pump  220 , and a motor  230  for driving pump  220 . A reservoir  240  connected to an air bag  250  with a conduit and a damper  260  parallel with the expansion axis of the bag is further provided. A reservoir  2770  connected to an air bag  280  with a conduit, and a damper  290  parallel with the expansion axis of the bag is further provided. It is appreciated that the volume of the reservoirs compared to the distance the bags are inflated result in a near constant force element being provided to support the booms. The conduits between the reservoirs and bags can be located inside of other components or outside. The pump  220  can rapidly change direction causing the left or right side bags to selectably inflate or deflate (without changing he amount of gas in the system) and accordingly raise the left boom or right boom angularly relative via center rack positioning while the opposite boom is lowered angularly. 
     Height sensors  300  are preferably located at the root end of the primary section, the outboard end of the secondary section and the outboard end of the breakaway section on each side boom. Hence, it is preferred to have six sensors. Of course, the number and location of the sensors  300  could change without departing from the broad aspects of the present invention. A pressure sensor  310  can also be provided to measure pressure on each side of the center rack within the reservoirs. 
     As discussed with more particularity, it is seen that a center rack  170  incorporating four of these planar linkages  180 ,  181 ,  182  and  183 , one at each corner, is provided. Each of these planar linkage center struts are attached, via the spherical joint at the distal end of the center-strut, to the support frame connected to the vehicle by the conventional four-bar lift linkage. Thus it may be observed that if the center-rack&#39;s mass were to be suitably supported, then it would be free to move vertically, horizontally and in angular translation about the vehicle&#39;s longitudinal axis, within the movement limitations of the planar linkages, while being constrained in all other degrees of freedom. While four planar linkages are shown, the present invention is not limited to utilizing four such linkages. 
     The center rack  170  with booms attached, being restrained by the four planar linkages as described, above, but with the weight of the boom assembly being reacted by two resilient suspension elements. These suspension elements can be mechanical springs, gas struts, liquid springs, hydraulic struts connected to compressed gas accumulators to act as springs, air springs, Near Constant Force (NCF) elements or any other type of resilient element that would reduce the shock loads that would otherwise be imposed on the boom assembly by vertical or roll displacements of the vehicle. These two spring elements may also be mounted in parallel with dampers, in much the same way that car suspension springs and dampers are configured to prevent resonance of the suspended mass (the center-rack and booms) by converting resonance momentum into heat, and dissipating it to their surroundings. 
     In one specific embodiment of the invention, the spring elements are gas springs configured to act as low K springs, that is, having a very small increase in spring force with displacement. This is achieved by the entrapped volume of the springs being large in relation to the swept volume displaced for a given spring movement. These gas springs  250  and  280  may take the form of air-bags similar to those often used for heavy truck and trailer suspension systems, but to increase the displacement/volume ratio, an additional closed reservoir volume may be attached to the airbag via a large bore connecting pipe (to minimize flow losses). Configured in this way, the effective K value of air springs may be lowered to the point that the spring system may be regarded as a Near Constant Force (NCF) element. Very low spring rate and NCF elements are advantageous in isolating or decoupling the suspended mass of the boom system from spurious movements of vehicle, but since they have little or no convergent restoring force, they require some form of active control to keep the boom system with the planar linkages limits of operation. 
     It should be noted that any lateral forces that would otherwise cause the center-rack and booms to displace sideways under lateral accelerations imposed by the differential vertical movements of the wheels on either side of the vehicle, combined with the height of the boom and center-rack system above the ground, may be countered by spring elements and dampers, positioned to act laterally to restrain the center-rack relative to the center-rack support frame. The vertical position of these spring restraints may be constrained to be approximately coincident with the combined booms/center-rack assembly vertical center of mass in order to minimize spurious vertically acting forces from being imposed on the center-rack under lateral loadings or in side slope operation. In another variation of this embodiment of the invention, two airbag spring elements, instead of the single restraining spring are mounted at each side of the center-rack to restrain it. 
     Two airbag type low spring (K) rate or NCF elements supporting the suspended mass of the center-rack and boom system can be provided. Height level sensors are fitted to each side of the center-rack, so that the height of each end of the center rack relative to the center-rack&#39;s support frame can be determined; accordingly their two signals can be computed to provide a combined mean rack height position. That is to say that the mean height level can be known, irrespective of the any angular displacement in roll between the decoupled center rack and boom system and the vehicle mounted boom support frame. Also, the (typically) two actuators (usually hydraulic cylinders) that are fitted to the vehicle&#39;s four-bar linkage that raise and lower the center-rack and boom assemblies en-mass, are fitted with positional transducers (typically linear transducers) that enable the center-rack height position to be computed at any time. Further, the two inner booms are furnished with a multiplicity of ground proximity height sensors  300 , with at least one sensor at each end of each inner or primary boom section, to measure the height above the ground. Similarly, the secondary and breakaway boom sections are also fitted with ground height proximity sensors: At least one at, or towards the outboard end of the secondary boom section, or on the breakaway section, or both. 
     Additionally, either pressure sensors are fitted to each of the air bags to measure the dynamic air pressure within them, or force sensors are used to measure the dynamic force being applied or reacted by each of the two air bags into the boom structure. 
     The enclosed volumes of the two airbag springs are interconnected by a large bore tube and a high throughput, bi-directional, positive displacement air pump  220 , such as a Roots pump, is interposed in this interconnecting line, such that when driven in one direction of rotation, air is displaced from the left side airbag into the right side air bag, and when rotated in the opposite direction, air is pumped in the opposite direction, from the right side airbag into the left side air bag. 
     A further, smaller positive displacement leveling pump  210 , serves to pressurize the whole boom suspension system, while control valves are fitted to operate in conjunction with the leveling pump to increase or decrease the mean pressure within the enclosed volume of air within the airbags, interconnecting pipe and bi-directional positive displacement pump. 
     In operation, the entire boom system is supported by the two airbag springs, which are pressurized by the leveling pump to lift the center-rack and booms to the correct mean height to optimize the available movement of the four planar linkages. The leveling pump may be driven electrically from the vehicle&#39;s electrical supply, or driven by a hydraulic motor from the vehicles hydraulic supply. The leveling pump may exhaust into a pressurized air reservoir or accumulator from which the control system monitors and controls the pressurized air leveling supply to the enclosed boom suspension system by means of the valves mentioned earlier. Either way, the leveling operation may be controlled electronically using signals processed from positioning sensors mounted between the center-rack and the vehicle mounted support frame to control the pressurized airflow into and out of the enclosed system volume. In an alternative embodiment, a mechanically operated leveling valve, as used to self-level the ride-height of commercial vehicles may be employed. 
     Again, in operation, once the center hack height has met, and is being maintained at the required mean dynamic height position, then signals from the ground height sensors, mentioned in this specification in the section “Background to the Invention”, mounted on the boom inboard or primary boom sections, are used dynamically level the booms relative to the ground. This is done by electronically analyzing the relative inboard boom heights above the ground, generating the required (dynamic) set point and deviation signals, and driving the positive displacement pump in the appropriate direction to change the relative forces being monitored and applied to each of the airbags in order to permit the boom systems weight and mass to effectively drive the boom in angular displacement on an axis parallel to the vehicle&#39;s roll axis, to cause the boom system to roll, en-masse, towards the dynamic set point where the inner booms on both sides of the vehicle are at the same height from the ground. Of course, the inertia of the booms will typically cause the boom movement to pass the set point, whereupon the direction of rotation of the positive displacement pump will be reversed to correct, and the booms will continuously be “balanced” in this manner, back and forth, although almost imperceptibly, such that the inner boom sections will remain at equal height above the ground on both sides of the vehicle, irrespective of the undulations and topographical contour changes of the ground profile. 
     Now, achieving the dynamic balance of the inboard boom sections at equal heights above the ground is only part of the requirement of a boom system. If the vehicle is passing through a gully where the ground profile will rise up on either side of the vehicle, or along the top of a ridge where the ground profile will slope down on either side of the vehicle, simply having a straight boom span maintaining equal heights both sides of the vehicle, could prove inadequate. In the former condition, the gully, not only could the greater part of the booms be below optimal height for spraying, but the outboard sections of the booms could actually impact the ground: In the latter condition, the ridge, the outboard sections of the boom could be so far above the optimal spray height that the spraying operation could be almost ineffective. Accordingly, essentially simultaneously with the balancing of the complete boom and center-rack assembly to equal mean height above the ground on both sides of the vehicle, each of the two secondary boom sections is arranged to pivot in the region of their folding hinge point to the inner, or primary boom sections, upwards and downwards about longitudinally disposed pivot axes (in the operating mode): That is, that the primary and secondary boom sections can be articulated relatively to each other to increase or decrease their dihedral/anhedral angle as required to maintain the secondary at the correct mean height above the ground irrespective of the complex and varying span-wise contours of the ground, and the dihedral/anhedral positioning of the inboard boom sections which result from their own contour following capabilities. 
     Turning now to  FIGS. 12-14 , it is seen that a positional connector  350  is provided. The connector  350  connects a primary boom section  360  with longitudinal axis  361  to a secondary boom section  371  with longitudinal axis. A folding actuator  380  is shown schematically, which is used to fold the secondary boom section relative to the first boom section. A top pivot  390  and a bottom pivot  400  are provided. The top pivot is a pivotal connection of a fixed length. The bottom pivot  400  has an actuator  401  and a positional control  402 . The pivot is rotational as well as linearly adjustable. The actuator can have a predetermined stroke, wherein at one end of the stroke the secondary boom section  370  is held on an inclined plane or orientation relative to the primary boom section  360 . Yet, when the actuator  401  is at the other end of the stroke, the secondary boom section  370  is held at a declined plane or orientation relative to the primary boom section  360 . The secondary boom section  370  can further be oriented wherein its longitudinal axis  371  is generally parallel to the longitudinal axis  361  of the primary boom section at a point intermediate the two actuator stroke ends. It is appreciated that the actuator  401  can be controlled by the controller  402  to make adjustments in real time based on inputs from height sensors. Actuator  401  is preferably a hydraulic actuator. 
     Stated more particularly, in operation, the ground proximity sensor  300 , or sensors mounted at the outboard end of the secondary boom section and/or breakaway measures the height of the outboard end above the ground. The controller  200  compares this value with the height sensor at the outboard end of the inner or primary boom section, and a deviation signal generated. The controller in turn corrects the dihedral/anhedral angle of the outboard boom section and breakaway relative to the boom inboard section by controlling the hydraulic actuator between the two sections, such that outboard end of the secondary and or breakaway boom sections is brought to similar height above the ground to the outer end of the primary section. 
     Turning now to  FIG. 15 , it is seen that a high level flow diagram  420  of a control system for this particular embodiment of the invention.  FIG. 15  shows the generic type and location of the minimum number of sensors required to operate the boom leveling and control system. The right hand column shows the generic type of input sensors while the lower left column shows process actions with some feedback signals. The top configuration table is incorporated to account for significant changes in operational inertia data, such as operating the booms in their short span semi-folded configuration, or to permit the folding and unfolding of the booms under full control while the vehicle is moving. 
     In another embodiment of the current invention, any propensity of the boom whole boom system to resonance in the vertical or “flapping” mode, particularly at its Eigenfrequency may be countered by turning the boom system itself into a tuned mass damper, tuned to its own first-order natural frequency, This is achieved by arranging for the spring rate (K) of the combined airbags at the center-rack that support the mass of the boom/center-rack system, to have closely the same natural frequency as the boom span itself. The planar linkages allow the free vertical displacement of the center rack to achieve the necessary freedom of movement. The dampers that are in parallel with the airbags then serve to dissipate the energy of the resonance—in the manner and function of a tuned mass damper—to reduce the amplitude of resonance to a much lower and far less destructive level. The system may be tuned by proportioning the suspending airbag springs appropriately during the design to give the appropriate spring rate and/or adjusting the internal volumes of the air reservoirs attached to each airbag by means of an adjustable piston at the closed end of the reservoir, or by having a hydraulically displaceable diaphragm within the reservoir and pumping the oil entrapped behind the diaphragm in or out in order to vary the internal air volume of the reservoir. 
     This concept is very attractive, since the booms become effectively become their own tuned mass damper, but without having to add an additional mass which, of course would be necessary with the more conventional approach to tuned mass damping. Such systems can be readily modelled, indeed reduced to practice, by using Finite Element Analysis (FEA) to determine the Eigenfrequency of the vertical flapping mode of resonance of the boom system, and multi-body dynamics software, such as ADAMS, to model the TMD damping effect. 
     Turning now to  FIGS. 16-18 , it is seen that a preferred embodiment of a tension gas spring  450  is provided. The tension gas spring  450  has a housing  460  with two ends  461  and  462 . A slot  463  or other type of opening is provided spanning generally longitudinally along one or two sides of the housing  460 . A gas strut  470 , preferably a compressive gas strut with a damping components, is provided and is fixed at the top end  461  of the housing. The strut  470  has a first end  471  and a second end  472 . The first end  471  is preferably pinned to the housing  460  at or near the first housing end  461 . It is preferable that the seal is oriented down wherein the damping fluid can cover the seal when the unit is stored to preserve the integrity of the seal. The opposite end  472  of the actuator can move along a longitudinal axis relative to the housing. An arm  480  is pivotally connected to the second end  472  of the strut. A stirrup  491  can be connected to the end  471  with a bolt. A second arm  490  is pivotally connected to the second end  462  of the housing in a fixed longitudinal position. A stirrup  491  and bolt  492  are used to connect the arm  490  to the housing. It is understood that the stirrups are pivotally connected to the arms allowing for rotation there between. While stirrups are shown, it is appreciated that alternative connective structures could be used without departing from the broad aspects of the present invention. 
     In use, a force can be provided to force the arms away from each other (within the constraints of the gas strut or spring). There is preferably little or no damping provided in this direction. However, the tension spring  450  of the present invention applies a return force to the arms (biasing them towards each other) and provides an amount of damping at the latter end of the return stroke. The amount of damping can be determined by a variety of factors relative to the compressive gas shock used in tension spring. 
     Looking now at  FIGS. 19-21 , it is seen that a breakaway section tuned mass damper  500  is provided to damp boom resonance in the fore and aft flapping mode. The breakaway section tuned mass damper  500  has a secondary boom section  510 , a breakaway boom section  520  and two tension springs  530  and  540 . Spring  530  is on one side of the boom with one arm connected to each boom section. Spring  540  is on the opposite side of the boom and also has one arm connected to each boom section. 
     The breakaway section  520  is free two swing out laterally relative to the secondary section  510  in either direction without encumbrance. In this regard, the breakaway section can retain its intended function. Yet, the tension springs are used to bias the breakaway section to an orientation back in line with the secondary section ( FIG. 19 ). In addition to the biasing force provided by the springs  530  and  540 , the springs  530  and  540  provide dampening thereby turning the breakaway section into a tuned mass damper (without adding appreciable weight to the system). It is appreciated that depending upon the direction of the swing, that only one of the tension springs  530  or  540  is actively damping the system. 
     According to another embodiment of the current invention, one or more tuned mass dampers (TMDs) are in or on each boom semi-pan, at a position or positions calculated or measured to place them in fairly close proximity to the anti-nodal regions at the boom system Eigenfrequency and/or at anti-nodal positions of any problematical secondary or tertiary frequencies. Tuned mass dampers used for this purpose may be of the passive or active types and may be of a commercially available design, or designed specifically for the purpose. They may be attached externally to the boom&#39;s structure or mounted within it. The TMDs may be oriented to counter resonance occurring in a single plane, i.e., the horizontal plane to counter resonance in the vertical direction, or operate to counter resonance in more than one plane, i.e., to counter resonance in both the vertical and longitudinal planes. 
       FIG. 22  shows the principle of operation of just one of the many TMD configurations that may be used for the purpose of quelling resonant vibrations in spray booms. In this particular case the TMD is of the passive type and serves to operate to counter resonance in both the vertical and longitudinal directions, notwithstanding the Eigenfrequencies may be significantly different in these two resonant modes due to the boom&#39;s flexibility characteristics being bound by different structural and dimensional requirements in these two different orientations. 
     Referring again to  FIG. 22 , it is seen that a tuned mass damper  550  is provided having a base  560 , a mass  565 , a bar spring  570 , a first damping rod  570  and a second damping rod  575 . Component  565  is a substantial mass, supported on a rectangular cross-section beam bar spring  570 , which is in turn connected at its fixed end to substantial base  560 . The base  560  is bolted solidly to the boom structure though its four mounting holes at a boom span-wise location carefully calculated or empirically determined to render the TMD&#39;s function most effective. In positioning the TMD on or within the boom&#39;s structure, the elongate orientation of bar spring  570  is arranged to be essentially parallel to the elongate direction of the boom. The natural frequency of the mass  565  as it oscillates up and down, in essentially vertical displacement on bar spring  570 , when excited by vertical vibratory movements in the boom fed in through base  560 , is determined by the sectional characteristics of the bar spring  570 , and can be tuned to a specific frequency by varying the bar spring cross section horizontal width and vertical height, effective beam length and, of course the value of mass  565 . This can be similarly achieved for the natural oscillating frequency in the lateral direction. The TMD will be most effective in quelling the resonant frequency of the boom in each of the two resonant directions, vertical and longitudinal (relative to the vehicle axes) when the natural frequency closely matches the booms resonant frequencies in each of these two different directions. 
     In order to function as an effective TMD, the TMD mass is appropriately damped and, in the TMD depicted in  FIG. 22 , this is achieved by the means of two flexible damping rods, specifically rod  575  for damping in the vertical vibration direction and rod  580  for damping in the lateral vibration direction. These damping rods have their fixed ends attached at base  560 , and their free ends that are slidably constrained in tubes embedded in mass  565 . Accordingly, when the mass  565  oscillates in a vertical direction the free end of damping rod  575  is displaced slidably within its respective damping tube in mass  565 , such that viscous shear is induced in a damping medium such as thick silicone grease present within the tube and kept in place by a seal or a flexible bellows serving the same function. A similar damping effect is realized in the lateral oscillation direction by damping rod  580  moving within its respective tube in mass  565 , and again having a suitable damping medium sealed in by a seal. 
     It should be understood that this is a TMD concept advanced for explanatory purposes only. There are numerous ways a TMD can be configured, either as a passive device (as shown in  FIG. 22 , or as an active device typically using closed loop control and computer interfacing. A TMD can also be configured to impose minimal spurious reactive moments on the structure: For example, the device depicted in  FIG. 22  could be mirrored about the attachment face base  560  to provide a double mass device, reminiscent of a dumbbell, which would not introduce unnecessary vibratory bending moments into the supporting structure. 
     An embodiment of this nature is illustrated in  FIG. 23 . The tuned mass damper  600  in  FIG. 23  has a base  605  and a bar spring  610 . A first section  620  having a mass  621  and two damping rods  622  and  623  are provided. A second section  630  is also provided and is opposite of the first section  620 . The second section  630  similarly has a mass  631  and two damping rods  632  and  633 . 
     Two masses  621  and  631  are shown mounted at either end of a flexible bar spring  610 , which is itself supported at its central common nodal position by base  605  at its Eigenfrequencies in resonance in both the lateral and vertical modes, which effectively divide it into two sections, a first section  620 , and a second section,  630 . The supporting base  605  is, in turn, securely connected to the boom structure at, or close to, an anti-nodal position pertaining to the boom&#39;s resonant frequency that is to be damped. Damping rods  622  and  623  connect to hydraulic kinetic fluidic dampers to damp the first section,  620 , and damping rods  632  and  633  act similarly for the second section  630 . When the boom structure is excited towards resonance, in the horizontal or vertical modes, by imposed acceleration associated with the vehicle traversing rough terrain, or maneuvering terrain, the TMD responds by resonating at the same frequencies but out of phase with the boom structure, while the damping of the TMD acts to damp out both the TMD and boom resonance, dissipating the resonant energy as heat in the TMD&#39;s damper systems. 
     Similarly, damping of the oscillating mass or masses can be carried out by a variety of means other than by the viscous damping shown in  FIG. 22  to the kinetic fluidic damping described in  FIG. 23 . For example, friction (coulomb) damping, magnetic, electromagnetic or other damping methods can readily be employed. 
       FIGS. 24 and 24A  show a classical passive linear TMD which, in this embodiment of the invention, may be incorporated into a spray boom&#39;s structure to mitigate damaging resonance. The tuned mass damper  650  has a housing  660  with two ends  661  and  662  as well as a cover  663 . The cover  663  is shown in breakaway view so that the internal components can be readily viewed. A slider rod  670 , a mass  675  operable on the slider rod  670 , and two springs  680  and  685  are provided. The housing  660  can be filled with an amount of fluid  690  that provides damping to the tuned mass damper  650 . 
     Stated with more particularity, the mass  675  is freely slidably mounted on slider rod  670 , and constrained by spring  680  on one side and spring  685  on the other. The mass  675 , slider rod  670  and the springs  680  and  685  are themselves contained in a housing  660 , which has ends  661  and  662  which directly support slider rod  670  at its outer ends, and also act as end restraints for springs  680  and  685  at their outer ends, while the same springs serve to restrain the mass  675  at their inner ends. Thus, inertial movement of the mass  675  in sliding motion on slider rod  670  against one or other of the springs, can be defined mathematically in terms of resonant frequency movement, if undamped. However, since housing cover  663  covers housing  660  and the spring—mass—slider system, to seal it hermetically, while entrapping within the enclosed volume an amount of damping fluid  690 , the whole system becomes an effective TMD. The damping fluid may be a gas such as air, or a liquid. As a liquid, it may full fill or only partially fill the entrapped volume. 
       FIGS. 25 and 25A  illustrate another example of an embodiment of the current invention that embraces and demonstrates two principles: That of electromagnetic damping and/or active TMD control. A tuned mass damper  700  has a housing  710  with two ends  711  and  712  and a cover  713 . The cover is shown in breakaway view (with hatching) so that the interior components can be illustrated. The cover is mated with and has a diameter similar to the round ends of the housing. A slider rod  720 , a mass  725  operable thereon, two springs  730  and  735  and a coil  740  are provided. The inductive coil  740  entirely encapsulates the housing around the full perimeter of the mass  725 . 
     Stated more particularly, the housing  710  having ends  711  and  712 , and containing slider rod  720 , mass  725 , and springs  730  and  735 , is similar to that depicted in  FIG. 24 . However the housing,  710 , (shown sectioned to show the internal components) supports an inductive coil  740  (also sectioned for the same reason), that serves the function of damping. This is achieved in conjunction with mass  750  being constructed of magnetic material and being magnetized, or containing within it a magnet or magnets of the permanent or electromagnetic types. The linear bearing surface between the mass  725  and the slider may be of the plain bearing type, or of the gas-bearing or recirculating rolling element types to minimize wear and/or reduce friction. The outer cover  713 , may or may not be hermetically sealed, but does not contain a damping liquid. Typically the medium within the housing  710  would be ambient air, vented to minimize damping, or a sealed housing enabling the mass  725  to move freely in a vacuum for example. The primary source of damping for this form of TMD is caused by the electromagnetic inductance generated between the magnetic mass  725  and the inductive coil  740 . Since this is controllable between effectively shorting-out the coil output and dissipating the resonant energy as heat within the inductive coil, or dissipating the energy elsewhere using as part of a comprehensive control strategy, this methodology is defined as an active TMD. 
     Turning now to  FIGS. 26 to 28 , it is seen that the primary and secondary boom sections are readily adjustable by the structures and methods of the present invention. The left and right primary booms can conventionally be angularly adjusted relative to the center rack with hydraulic cylinders. Then, the center rack can be adjusted in a plane or near plane relative to a support assembly. By shifting air to one side of the other, the center rack can twist relative to the support frame while remaining in a same plane and generally parallel to the to support frame. Then, there is preferably a positional connector on each boom, wherein the secondary sections can be angularly adjusted relative to the primary sections. Hence, in the illustrated embodiments, it is seen the present invention can be used to create a boom pair that can maintain a desired spray height. 
     Thus it is apparent that there has been provided, in accordance with the invention, a planar linkage, methods of decoupling, mitigating shock and resonance, and controlling agricultural spray booms mounted on ground vehicles that fully satisfies the objects, aims and advantages as set forth above. While the invention has been described in conjunction with specific embodiments thereof, it is evident that many alternatives, modifications, and variations will be apparent to those skilled in the art in light of the foregoing description. Accordingly, it is intended to embrace all such alternatives, modifications, and variations as fall within the spirit and broad scope of the appended claims.