US 3923412 A
A vehicle mounted vibratory compactor employing a pair of vibrating elements that extend laterally across the entire width of the vehicle and are mounted, as a unit, in a rigid support frame. The frame is designed as a parallelogram with parallel control arms for resisting the bending forces applied thereto by the action of the vibrating elements compressing materials, such as soils of all types, and the reactive forces of the materials. The magnitude of the compression forces exerted by the vibrating elements is varied considerably by the utilization of unique drive means comprising diverse combinations of springs, reciprocal plates, pressure limit valves, pneumatic servomechanisms, pistons, and the like, for determining the natural frequency of the compactor system and thus controlling the resonant frequencies at which the maximum compressive forces are produced. The vibrating elements, which may assume the form of plates, rollers, beams or shoes, are fabricated from a plurality of simplified components.
Claims available in
Description (OCR text may contain errors)
United States Patent Linz 1 1 Dec. 2, 1975 1 1 DRIVE MEANS FOR VEHICLE MOUNTED 3.756.735 9/1973 Linz 404 133 VIBRATORY COMPACTOR 3,802,791 4/1974 Uebel 404/133  Inventor: Albert Linz, .lulweg 40,
Hoffnungsthal near Cologne, Germany  Filed: Sept. 9, 1974  Appl. No.: 504,560
Related US. Application Data Primary Examiner-Nile C. Byers, Jr. Attorney, Agent, or FirmAndrew 1. Beck; Stephen P. Malak; Martin P. Hoffman  ABSTRACT A vehicle mounted vibratory compactor employing a pair of vibrating elements that extend laterally across the entire width of the vehicle and are mounted, as a unit, in a rigid support frame. The frame is designed as a parallelogram with parallel control arms for resisting the bending forces applied thereto by the action of the vibrating elements compressing materials, such as soils of all types, and the reactive forces of the materials. The magnitude of the compression forces exerted by the vibrating elements is varied considerably by the utilization of unique drive means comprising diverse combinations of springs, reciprocal plates, pressure limit valves, pneumatic servomechanisms, pistons, and the like, for determining the natural frequency of the compactor system and thus controlling the resonant frequencies at which the maximum compressive forces are produced. The vibrating elements, which may assume the form of plates, rollers, beams or shoes, are fabricated from a plurality of simplified components.
15 Claims, 34 Drawing Figures //////y/////////// I04 I03 US. Patent Dec. 2, 1975 Sheet 1 of 14 3,923,412
F/GIZ US. Patent Dec. 2, 1975 Sheet 2 of 14 3,923,412
US. Patent Dec. 2, 1975 Sheet4 0f 14 3,923,412
U.S,. Patent Dec. 2, 1975 Sheet 6 of 14 3,923,412
US. Patent Dec. 2, 1975 Sheet 8 of 14 3,923,412
US. Patent Dec. 2, 1975 Sheet 10 0f 14 3,923,412
Patant Dec. 2,
Sheet 12 of 14 US. Patent Dec. 2, 1975 Sheet 13 of 14 3,923,412
v/2l4 r FIG. 30 2|2- FIG. 3/
US. Patent Dec. 2, 1975 Sheet 14 0f 14 3,923,412
DRIVE MEANS FOR VEHICLE MOUNTED VIBRATORY COMPACTOR This is a continuation of application Ser. No. 178,567, filed Sept. 8, 1971, which has now matured into US. Pat. No. 3,834,827.
BACKGROUND OF THE INVENTION 1. Field of the Invention The invention relates generally to vehicle mounted vibrator compactors, and more particularly to drive means for efficiently driving the pair of vibrating elements utilized by such compactors against the material to be compacted in an alternating fashion.
2. Description of the Prior Art Vibratory compactors of various designs have been widely used for compacting particulate materials, such as sand, gravel, soils of all kinds and consistencies, by repeatedly and rapidly applying vibratory compacting forces thereto. One variety of conventional vibratory compactors has included a vehicle frame, a motor for driving the frame to the work site, a transverse shaft supported on the vehicle frame, an array of heavy tamping shoes suspended from the shaft at spaced intervals, and take-off belts for delivering vibratory energy from the motor to each of the tamping shoes in sequence. Such multishoe vibratory compactors are disclosed in several United States patents, including US. Pat. No. 2,938,438, granted to W. L. Hamilton; US. Pat. No. 2,958,268, granted to Moir; and US Pat. No. 3,181,442, granted to J. H. Brigel.
Several problems have been encountered with the multishoe vibratory compactors; for example, alignment between the adjacent shoes has to be carefully maintained, lest the shoes collide with attendant damage to the shoes and impairment of compacting efficiency. Also, the spacing between shoes leads to a tendency to create a series of irregularities in the compacted material disposed therebetween; and the repeated raising and lowering, and vibrating movement, of the tamping shoes leads to excessive wear at the supporting bearings and/or the drive belts that deliver the vibrating forces to the shoes.
SUMMARY OF THE INVENTION The present invention overcomes the above-noted structural shortcomings and functional deficiencies of known vibratory compactors by providing a vehicle mounted compactor of simplified design that is characterized by a rigid parallelogram frame for supporting the vibrating elements, such frame being vertically adjustable within guides defined on the vehicles. Consequently, the vibrating elements can be elevated to pass freely over obstacles in the road bed as the compactor is driven to the work site, and the vibrating elements can be lowered upon reaching the site. Additionally, the amplitude of the stroke of the vibrating elements can easily be varied. The pair of vibrating elements, which may be plates, beam, shoes, or rollers, are supported by the parallelogram frame for both pivotal and vertical movement. The driving means for imparting vibratory energy to the vibrating elements can be operated over a wide range of frequencies, thereby enabling the compactor to satisfactorily compress a wide variety of materials. Vibration dampeners deaden harmful, reactive forces and extend the useful life of the instant vibratory compactor.
It is another objective to provide at least one embodiment of the rigid parallelogram support frame lacking a natural or inherent frequency, so that the vibrations imparted thereto by the driving mechanism are transmitted with minimal losses to the vibrating elements. In other embodiments, sundry combinations of springs are utilized to establish a desired natural, or inherent, frequency for the frame. A vibration generator is positioned directly on each vibrating element, so that the driving mechanism is far simpler in design than the conventional complex power takeoff belt arrangement utilized for sequentially driving each vibrating shoe in the multishoe compactors described above. The driving mechanism includes two cranks operating out of phase so that one vibrating element is engaged with the material to be compacted while the other vibrating element is raised to its highest point of travel.
It is another objective to provide spring means operatively associated with the lower end of the rods to transmit vibratory energy from the driving mechanism to the vibrating elements which alternately engage the material to be compacted.
It is yet another objective to provide, in certain embodiments of the vibratory compactor, series or parallel pneumatic circuits interposed between the cranks and the vibrating elements to function in the same manner as mechanical springs to influence the frequency of the compacting system. Such circuits, which include various combinations of pistons, manually adjustable pressure limiting valves, reservoirs, etc. may be used in lieu of mechanical springs, or in combination therewith.
It is'yet another objective to provide series or parallel pneumatic circuits, which function in the same manner as mechanical springs interposed between the cranks and the compacting elements. Additionally, the pneumatic circuits can be utilized in combination with the mechanical springs. The pneumatic circuits may include various combinations of pistons, manually adjustable pressure limiting valves, pneumatic reservoirs or accumulators.
It is still another objective to provide compensating weights located vertically above the vibrating elements and driven by the same cranks that drive the vibrating elements, but always 180 out of phase therewith. Such compensating weights dampen. the inertial forces reacting upon the frame of the vibratory compactor, and permit the utilization of a vehicle and frame of reduced tonnage without a reduction in operational efficiency.
It is still another objective to provide sundry combinations of compensating weights, discs, pneumatic servo-pistons, mechanical springs and the like, which transmit the destructive reactive forces set up by the vibrating elements compacting the soil to the chassis of the vehicle, rather than to the drive means therefor. Consequently, the life of the drive means is prolonged and the effectiveness of the vibratory compactor is enhanced.
Since the inertial reactive forces acting upon the vibrating elements are substantially reduced by virtue of the various unique constructional features noted above, the vibrating elements, per se, may be assembled from simplified components. The vibrating elements may thus be readily altered in size, and their light weight reduces the magnitude of their inertial forces. Sheepsfoot protuberances may be formed on the compacting surface. In contrast thereto, conventional vibrating shoes were heavy unitary members that lacked versatility in size and shape.
Other objectives and features of the instant invention will be pointed out in the following description and illustrated in the accompanying drawings, which disclose, by way of example, the principles of the instant invention, and the best modes contemplated for achieving such principles.
BRIEF DESCRIPTION OF THE DRAWINGS FIG. 1 is a side elevational view ofa vehicle mounted vibratory compactor constructed in accordance with the principles of the instant invention;
FIG. 2 is a horizontal cross-sectional view of the vibratory compactor, such view being taken along line II-II in FIG. 1 and in the direction indicated;
FIG. 3 is a vertical cross-sectional view of the vibratory compactor with the vibrating elements in the elevated position, such view being taken along line III-III in FIG. 1 and in the direction indicated;
FIG. 4 is a vertical cross-sectional view identical to the view of FIG. 3, but showing the vibrating elements in a lowered position;
FIG. 5 is a side elevational view, on an enlarged scale, of a preferred embodiment of the frame for supporting the vibrating elements;
FIG. 6 is a side elevational view of a modification of the vibratory compactor of FIGS. 1-5, such view showing additional tools mounted on the vehicle frame;
FIG. 7 is a combined side elevational view and schematic representation, on an enlarged scale, of a first alternative embodiment of the frame for supporting the vibrating elements and the drive mechanism for imparting vibratory energy thereto;
FIG. 8 is a horizontal cross-sectional view of the first alternative embodiment of the frame for supporting the vibrating elements, such view being taken along line VIII-XIII in FIG. 7 and in the direction indicated;
FIG. 9 is a graphical representation of the cycle of operation for the frame and vibrating elements shown in FIGS. 7 and 8;
FIGS. 10 and 11 are side and front elevational views, respectively, of the drive mechanism for imparting vibratory energy to one of the vibrating elements of the embodiments of FIGS. 1-5, 6, or 7-8;
FIG. 12 is a schematic representation of an alternative drive mechanism for imparting vibratory energy to one of the vibrating elements of the embodiments of FIGS. l-5, 6 or 7-8;
FIG. 13 is a combined side elevational view and schematic representation, on an enlarged scale, of a second alternative embodiment of the frame 'and the drive mechanism for imparting vibratory energy thereto;
FIG. 14 is a vertical cross-sectional view of the embodiment of FIG. 13, such view being taken along line XIVXIV and in the direction indicated;
FIGS. 15-19 are schematic representations of various mechanical spring systems, including appropriate guides, for imparting vibratory energy to the vibrating elements;
FIGS. and 21 are side and front elevational views, respectively, of another mechanical spring system for imparting vibratory energy to the vibrating elements;
FIGS. 22-25 are schematic views of pneumatic systems for imparting vibratory energy to the vibrating elements;
FIGS. 26 and 27 are side and front elevational views, respectively, of an alternative mechanical spring system for imparting vibratory energy to the vibrating elements;
FIG. 28 is a schematic view of a mechanical spring and compensating weight system for dampening vibrations;
FIG. 29 is a detail view of the compensating weight used in the system of FIG.28; and
FIGS. 30-34 show constructional details of the vibrating elements.
DESCRIPTION OF THE PREFERRED EMBODIMENTS Referring now to the drawings wherein similar reference numerals identify similar components, the preferred embodiment of FIGS. l-5, and the variant shown in FIG. 6, depict a carrier vehicle including a chassis 1 with two axles 3. Five pneumatic tires 2 are arranged on the axles 3 in an offset fashion so that the displaced tracks exert a compacting pressure upon the surface of the road or upon the material to be compacted. The carrier vehicle can be propelled forwards and backwards at variable speeds; and can be steered with great facility.
Vertical guides 4 are defined on opposite sides of the central portion of chassis l, and lateral extensions on the frame 5 for the vibrating elements 6 fit within the guides, as shown in FIGS. 2-4. Frame 5 may be supported for movement within the guides by rollers or guides, and the frame may be moved vertically by various conventional devices, such as worm gears, hydraulic actuators, etc.
In the preferred embodiment of FIGS. 1-5, the vibrating elements consist of a pair of vibrating plates, although vibrating beams, shoes, or rollers, can be substituted thereforI The constructional details of the frame 5 which supports the pair of vibrating elements 6 are best shown in FIG. 5. The vibrating elements are held parallel to one another by a pair of horizontally extending control arms 7; the upper control arm is secured to the end plate 8 by a pin or shaft 9, while the lower control arm is secured to plate 8 by a pin or shaft 10. The ends of the control arms 7 furthest from shafts 9 and 10 are secured to side supports 13 by bolts or shafts l4, l4, and corresponding slots function as hinges for the parallelogram frame 5 and permit a slight rocking movement of the vibrating elements in response to the vibration generators. The midpoints of control arms 7 pivot about shafts 9 and 10, which are situated in the same vertical plane passing through the center of plate 8. Roller bearings may be positioned about shafts 9, 10 to facilitate movement of the arms about the shafts.
Braces l 1, 12, seen in FIG. 5, extend between control arms 7 to further enhance the strength of the frame supporting the vibrating elements. The optimum location for the total center of mass for the means for imparting vibratory energy to vibrating elements 6 is located in the vertical plane defined by shafts 9 and 10.
The equilibrium position for parallelogram frame 5 is shown in FIG. 5. In the'equilibrium position wherein the bottom surfaces of the vibrating elements lie in the same horizontal plane, the shafts 14 are situated a short vertical distance above shaft 10, and shafts 14' are situated a short vertical distance above shaft 9. Accordingly, when one of the vibrating elements is lowered to its operative position, the frame will exert an unbalanc ing force with respect to the lowered element.
Vibration generators of conventional design are secured directly upon compacting elements 6, as shown in FIG. 5. The main engine (not shown) for the compactor, which may be an internal combustion engine, transmits power to the generators over a system of transmission belts, including belt 21, which is entrained over pulley on shaft 19, and horizontally extending power take-off belts 30, which is also entrained over pulleys 17 and 20. Shaft 19 extends laterally between cover plates 8 of frame 5, and is situated vertically intermediate shafts 9 and 10. The right hand vibration generator operates 180 out of phase with the force transmitted to the left hand vibration generator. Accordingly, the vibrating elements 6 will always be moving in opposite vertical directions; by virtue of the configuration of frame 5, the vibrating elements will also always be pivoting or oscillating about shafts 9, 10 in opposite directions.
In operation, vehicle 1 is driven to the work site with frame 5 and vibrating elements 6 in the elevated position shown in FIG. 3. At the work site, the frame and the vibrating elements will be lowered into the position shown in FIG. 4. Then, the elements 6 will press against the ground and be vibrated by vibration generators 15. Frame 5 can be vertically adjusted within guides 4 to accurately vary the amplitude, or length of travel to an extreme up or down position, from a mean position of vibrating elements 6. The oscillatingfrequency of the vibrating elements is adjustable by altering the rotational velocity of vibration generators 15. Furthermore, the weight of the vehicle exerted upon the offset pneumatic tires 2 as the vehicle moves to the work site enhances the efficiency of the compactor.
In the embodiment of FIG. 6, the carrier vehicle 1 is provided with oppositely disposed scraper blades 38, 39 and a scarifier 40. Consequently, such embodiment may be utilized as a road working machine ofincreased versatility for handling bituminous or similar paving materials along a roadway.
The alternative embodiment of FIGS. 7-8, shows a frame, indicated generally by reference numeral 102, and including large plates 102a at opposite sides of the chassis mounted for movement relative to the carrier vehicle 101. Two vibrating elements 103 are secured to each end plate 102 and large plate 102a of the frame by shafts 104 which pass through the horizontally extending control arms 104. A crank 106 and a rod 107 are connected to the end of the control arm, and a spring 105 extends between the lower end of rod 107 and vibrating element 103 and opposes the downward movement of the rod. Springs 105 may be spiral springs or torsion springs, and crank 106 and rod 107 are only schematic representations of one type of conventional vibration generator.
In operation, cranks 106 are driven from the prime mover in a conventional fashion, so that the elements 106 start to vibrate with a phase lag of 180 between one another. The frame and the vibrating elements, because they are influenced by springs 105, are a part of an oscillating or vibratory system that has a definite natural frequency. The amplitude of vibrating elements 103 is usually very small but increases sharply when the speed, or frequency, of the crank rotation falls within the resonance range for the springs.
To illustrate, assuming that frame 102 is lowered within guides defined in the body of vehicle 101 to a position wherein the compacting elements are 2 centimeters above the ground, and further assuming that crank 106 starts to rotate, at a speed or frequency half, or less,'than the natural frequency of the system, then the vibrating elements will be subjected to a forced oscillation of such a minute amplitude that they will not even completely close the 2 centimeter gap and touch ground. However, as the frequency of cranks 106 approach the natural frequency of the system, the driving forces imparted by the prime mover to the vibration generators 15 increase in effectiveness. Finally, when the frequency of cranks 106 equals the natural frequency of springs 105, all of the energy imparted by vibration generators 15 is effectively delivered to vibrating elements 103. The matching or coincidence of the driving frequency of the cranks to the natural frequency of the system is known as resonance magnification, and the compacting process begins with the attainment of resonance magnification.
FIG. 9 graphically represents the course of the oscillatory movement of vibrating element 103 during a cycle of operation. Assuming that the movement of element 103 approximates a sinusoidal path of travel and plotting the amplitude of movement versus time, A, designates its upper amplitude, A,, its lower amplitude, and T the period of time for execution of a complete cycle. T is the time interval during which the vibrating element compresses the ground and performs work, T is the interval during which the vibrating element releases its compressive forces and starts its upward movement away from the ground, and T is the interval during which the vibrating element does not touch the ground at all.
As noted previously, vibrating element 103 only approximates a sinusoidal function. The variations from the theoretical sinusoidal function are attributable, in part, to the effect of the natural or inherent frequency of springs 10S and the mass of frame 102 and the components suspended therefrom. Also, the movement of the oscillatory system changes rapidly at time T, because of the impact of vibrating element 103 with the ground, and the reaction forces thereby set up in the vibrating element and the frame supporting same. During the time interval T during which the element 103 is free of the ground, the path of oscillatory movement closely approximates the ideal or theoretical function f, for the companion element 103 only exerts a minimal influence upon the system.
During the time interval T the vibrating element accomplishes work compressing the plastic, semiplastic, or hard soil, or other particulate material, while overcoming the initial inertial resistance of the frame to movement. During upward movement, the vibrating element may be assisted in small measure by the resiliency of the material being compacted. However, in most instances, the resiliency of the material being compacted becomes an insignificant factor when compared to the amplitude of the vibrating element. Also, when the compacting forces applied to the material are increased significantly by resonance magnification, the magnitude of the resilient force of the material becomes insignificant by comparison.
While the embodiment of FIGS. 7-8 utilizes a single spring 105 for each vibrating element, the embodiment of FIGS. 10 and 11 utilizes a pair of coil springs 108 in lieu of a single spring. Both springs are situated within cylinder 109 which tapers outwardly at its opposite ends, and the rod 107 is driven by crank 106 and is secured to the upper end of the cylinder. Two diametrically opposite, axially extending sections 110 are removed from the central portion of cylinder 109, and stub shafts 115 extend through open sections 110 and maintain a spacer 111 intermediate the two springs. One end of each stub shaft extends axially through sleeve 112 which concentrically surrounds cylinder 109, while the opposite end of each shaft is mounted for pivotal movement by a journal 116 in bracket 113. The brackets are fastened directly to the upper surface of vibrating element 103.
.The pair of springs 108 are easily assembled within cylinder 109 in operative relationship in the following manner. The shafts 115 are removed from engagement with brackets 113 and end caps 114 are removed from the opposite enlarged ends of the cylinder. Springs 108 are then inserted into the cylinder, and end caps 114 are then pushed into position, firmly seating the springs. Aligned holes are drilled through the central portion of the first end cap, the spacer 111 and the second end cap. One end of rod 107 is then seated within the hole in the upper end cap.
Due to the oscillatory movement of crank 106, rod 107 moves vertically and the upper end cap 114 slides within the enlarged upper end of cylinder 109. The cylinder also pivots about stub shafts 115 in response to the movement of rod 107. The movement of the end cap compresses the upper spring 108, and the resultant forces are transmitted through spacer 111 and stub shafts 115 to brackets 113 and thence to vibrating elements 103. The springs are retained in fixed position by annular projections 117 on end caps 114 and annular projections 118 on spacer 111. The cylinder 109 prevents kinking of the springs as they are compressed.
During the cycle of operation of the system embodying the elements shown in FIGS. 7 and 8, the natural frequency of the oscillating system is determined by the combined mass of pivotally mounted cylinders 109 and the springs 108 seated therein.
The embodiment of FIG. 12 utilizes a pair of spaced parallel springs 119 in lieu of the springs 105 of the embodiment of FIGS. 7 and 8, or in lieu of the opposed pairs of prestressed springs 108 of the embodiment of FIGS. -1 1. The rod 107, which is driven by the crank at one end, is secured to a beam 120 at its opposite end. The parallel springs 119 extend between the beam and the vibrating element 103, and beam 120 is movable vertically within guides in frame 102. Both springs act as pressure springs, and together with the other components secured to frame 102, form an oscillatory system. Such system may have the same natural frequency as the oscillating systems of FIGS. 7-9 and 10-11, or it may have a different natural frequency.
In the embodiment of FIGS. 13 and 14, pressure springs 130 extend through cylinders 130' at opposite ends of vibrating element 103, and pressure spring 131 is connected between plate 132 and the midpoint of element 103. Plate 132 is connected to the lower end of rod 107, which is driven by crank 106. The upper ends of springs 130 rest against ledges 130" on the inner surface of frame 102. Springs 130 and 131 have the same effect upon the natural frequency of the oscillatory system as springs 105 in the embodiment of FIGS. 78. Under normal operating conditions, a portion of the reactive forces caused by the vibrations of element 103 striking the ground is transferred through springs directly to the frame 102, while by-passing the drive mechanism 106, 107, 131 and 132. At resonance, the ratio of the reactive forces transferred directly to the frame 102 to the reactive forces transferred back to the oscillation drive mechanism, is directly proportional to the ratio of the force of springs 130 to that of spring 131. Phrased in another manner, the larger the ratio of the force of springs 130 to that of spring 131, the greater the amount of reactive force transferred directly to the frame 102. The frame, obviously, is of sturdy construction and can easily withstand forces of such magnitude thereby increasing the operational life span of the other components of the drive system.
Additionally, the pair of vibrating elements 103 are supported for vertical movement, as well as pivotal movement, by a rigid parallelogram support frame including spaced side supports, a pair of control arms 104, and a pair of spaced shafts 104 about which the control arms can pivot.
FIG. 15 is a detailed view of the vertical guide 133 that controls the movement of beam 132, which is connected to the lower end of rod 107. Springs 130 and 131 are also provided with suitable guides, such as cylinders 133 to prevent sidewards kinking during compression thereof.
FIG. 16 illustrates another guide adapted to restrain the movement of two, or more, springs. The guide, which is indicated generally by the reference numeral 133a, includes a pair of rods 134 extending upwardly from vibrating element 103, and a beam 135 which moves vertically along both rods. Rods 134 extend upwardly through guide surfaces 136 which are secured to the guide 133a; the guide, in turn, is secured directly to the frame 102. A first spring 138 is located on each rod between the upper end of compacting element 103 and beam 135; a second spring 139 is located on each rod between beam 135 and guide surface 136; and a third spring 140 is located on each rod between guide surface 136 and enlarged stop 137 atop the rod. All of the springs are prestressed in the position shown in FIG. 16.
When crank 106 is rotating in a clockwise direction, springs 139 are compressed while springs 138 are relaxed, and vibrating element 103 follows the vertical movement of beam 135 due to the action of springs 140 and guide 133a. At the midpoint in the cycle the process is reversed, and springs 138 are compressed while springs 139 are relaxed, so that vibrating element 103 follows the movement of beam 135 against the resistance of springs 140. The vibrating element 103 oscillates due to the vertical reciprocation of beam 135. When the resonance frequency of the system is reached, the system is damped only by air resistance, internal friction in springs 138, 139 and 140, and by the work done during compacting the ground.
FIG. 17 illustrates another guide adapted to restrain the movement of three parallel springs. The guide, which is against indicated generally by reference character 133b, includes an additional spring 141 that is positioned about a rod 142 secured to the upper surface of element 103. Spring 141 extends between a stop 143 at the upper end of rod 142 and a supporting surface on the cylindrical member 144 that is mounted beneath plate 145. The cylindrical member is concentric with rod 142. Consequently, as plate 145 is vertically reciprocated by crank 106 and rod 107, the movement of cylindrical member 144 compresses, and releases, spring 141. In all other essential details, guide 133k functions in the same manner as guide 133 in FIG. 16.
FIG. 18 depicts another mechanical spring system for effectively transmitting vibratory energy from an internal combustion engine, or other conventional prime mover, to the vibration generator, and thence to the vibrating elements 103. A beam 155 is secured across the open upper end of element 103, so that the beam and the interior walls of the element define a chamber 148 therebetween. An aperture is formed in the center of the beam, and a piston rod 150 with an enlarged lower end 149 extends therethrough. A first spring 146 extends between end 149 and the bottom surface of chamber 148. A pair of spaced rods 134 are secured to the upper surface of beam 155, and a pair of guide surfaces 136 extend inwardly from guide 133. A first spring 139 is located on each rod 134 between beam 155 and surface 136, and a second spring 140 is located on eachrod 134 between surface 136 and the enlarged stop 137 atop rod 134. During a portion of the cycle of operation of the vibratory energy transmitting system, crank 106 and rod 107, beam 155 and thus vibratory element 103 is influenced by the compression of spring 146 and 139 while the springs 140 and 147 are relaxed. After passing the midpoint in its cycle of operation, beam 155 is influenced by the compression of spring 147 and 140 while springs 139 and 146 are relaxed.
FIG. 19 shows a variant of the mechanical spring system of FIG. 18 wherein the chamber 148 defined between beam 155 and the internal walls of vibrating element 103 has been enlarged. Also the springs 146, 147 operatively associated with piston rod 150 have been eliminated, rods 134 have been lengthened, and beam 150' is secured to the lower end of rod 150 and moves vertically along rods 134. Springs 160 are tensioned, when springs 161 are relaxed, and vice versa. Springs 160 act in concert with springs 139, while springs 161 act in concert with springs 140. The net result is that energy imparted by crank 106 and rod 107 is effectively transmitted by the springs and beam 150' to parallel beam 155 and vibrating element 103.
FIGS. and 21 disclose a mechanical spring system for transmitting a portion of the spring forces directly to the frame for supporting elements 103, such as frame 102 shown in the embodiment of FIGS. 7 and 8. Brackets 162 and 163 are secured directly to the large plate 1020 of frame 102. A first spring 164 has one end centered and seated firmly in bracket 162 and the rest of the spring extends downwardly toward bracket 163. A second spring 165 has one end centered and seated firmly in bracket 163, and the rest of the spring extends upwardly toward bracket 162. A bar 166 extends horizontally across the upper end of supports 167, and the lower end of spring 164 and the upper end of spring 165 are centered and seated firmly therein. The lower end of supports 167 is connected to one end of control rod 104. Alternatively, the lower end of the support could be connected directly to vibrating element 103.
If the maximum amplitude of the vibrating element is assumed to be 40 mm., and the total travel of the element is 80 mm., the length of the control rod is selected such that the maximum lateral deflection of springs 164, 165 amounts to only 0.8 mm. Thus, the danger of lateral kinking of the springs is obviated. FIG. 22 shows a pneumatic servo-piston that may be used in combination with mechanical spring systems shown in FIGS. 10-21, or in lieu thereof. Such pneumatic servo-piston may be used as a solitary control device or as an integral part of a pneumatic control circuit. The pneumatic servo-piston includes a cylindrical housing 168 and a piston 169 mounted on piston rod 170 so as to be movable within the housing. A first pressure limiting valve 171 is operatively associated with the mid-section of the housing, while a pair of pressure limiting valves 172 are situated at opposite ends of the housing. Valves 172 limit the maximum pressure built-up within the housing, while valve 171 determines the minimum pressure level within the housing. Consequently, the pneumatic servo-piston operates with a particular volumetric efficiency over a selected range of pressures, and closely approximates the action of a mechanical spring. These characteristics can be adjusted by varying the settings of valves 171 and 172 so that the power transmitted to vibrating elements 103 from crank 106 and rod 107 can be adjusted. Also, the valve settings may be altered to overcome the effects of inertia when starting crank 106 in motion or when attempting to stop same once it is in motion.
FIG. 23 shows a pneumatic circuit including the pneumatic servo-piston of FIG. 22 and further including a container 173 of pressurized fluid that is connected through pressure reducing valve 174 to reservoir 175. The pressurized fluid in container 173 influences the movement of piston 169 and piston rod 170 through valves 171 and 172. By utilizing the closed circuit of FIG. 23, the servo-piston operates independently of ambient air conditions and pressure changes. Additionally, the valves may be adjusted so that the servo-piston may be adjusted into harmony with the operation of the various mechanical spring systems.
FIG. 24 shows another pneumatic servo-piston that can be utilized within the power transmitting system of the instant invention. Such servo-piston includes a cylindrical housing 176 which may be open at both ends, and a middle partition 177 which extends perpendicular to the axial dimension of the cylinder and divides same into two chambers. A pair of pistons 178 are spaced along piston rod 179 at opposite sides of partition 177. The rod passes through an aperture in the center of partition 177, and the pistons move along the inner wall of housing 176.
FIG. 25 shows a pneumatic circuit including the pneumatic servo-piston of FIG. 24 and further including a container 191 of pressurized fluid that is connected through pressure reducing valve 184 to reservoir 195. The chamber formed above partition 177 is connected to container 191 via conduit and adjustable valve 192, while thechamber formed below partition 177 is connected to container 191 via conduit and adjustable valve 193.
Anequilibrium condition occurs during the operation of pistons 178 and piston rod 179 wherein the fluid entering housing 176 from container 191 on each piston rod stroke is equal to the fluid leaving housing 176 and returning to container 191. Such condition is related to the volumetric efficiency of the servo-piston, and influences the power transmitted to vibrating element 103. The equilibrium condition can be altered by varying the settings of valves i192, 193 and 194.
The motion of elements 103 can be stopped even though crank 106 and rod 107 are running by the simple expedient of venting valves 192, 193, to relieve any