US3131878A

US3131878A - Rock crushing apparatus with sonic wave action

Info

Publication number: US3131878A
Application number: US200091A
Authority: US
Inventors: Jr Albert G Bodine
Original assignee: Individual
Current assignee: Individual
Priority date: 1962-06-05
Filing date: 1962-06-05
Publication date: 1964-05-05
Anticipated expiration: 1981-05-05

Description

y 1964 A. e. BODINE, JR 3,131,878

ROCK CRUSHING APPARATUS WITH SONIC WAVE ACTION Filed June 5, 1962 3 Sheets-Sheet 1 INV EN TOR.

FIG.1

ATTORNEY ALBERT G. BODINE,J R.

May 5, 1964 A. G. BODINE, JR 3,131,878

ROCK CRUSHING APPARATUS WITH SONIC WAVE ACTION Filed June 5, 1962 3 Sheets-Sheet 2 INVENTOR.

ALBERT G. BODINE JR ATTnFeNEY y 1964 A. G. BODINE, JR 3,131,878

ROCK CRUSHING APPARATUS WITH some WAVE. ACTION 3 Sheets-Sheet 3 Filed June 5, 1962 INVENTOR.

ALBERT G.- BODINEQR.

l/ ATTORNEY United States Patent 3,131,878 RGCK CFtUSl- G APPARATUS Willi SGNEC WAVE ACTION Albert G. Bodine, Ilia, Los Angeles, Calif. (7877 Woodley Ava, Van Nuys, Calif.) Filed Zone 5, 1962, Ser. No. 2%,691 13 Claims. (CL Z41262) This invention relates generally to the processing of minerals and the like by gradual reduction of the size of relatively large particles thereof, and is especially applicable to rock crushing, though not limited thereto. The invention deals more particularly with crushing of particles of rock and the like by subjecting them to the fatigue action of powerful high impedance sound waves. In this connection, a sound wave of high impedance denotes a sound wave (wave of alternating compression and rarefaction) characterized by a high ratio of applied pressure wave amplitude to resulting elastic displacement velocity.

Conventional rock crushers commonly employ two jaws, one stationary and one movable, the movable jaw being forced toward the stationary jaw usually by a large toggle. Very extreme forces are required to crack and crush large rocks between the jaws of such Crushers, so that the crusher must be built ery large and heavy throughout, and must include very large and heavy foundation structure, in order to withstand the extreme stresses developed in the mechanism.

A primary object of the present invention is to provide a rock crushing process and apparatus operating upon novel sonic wave principles, and which, though provided with jaws of a fair degree of size and inertia, can, because of balancing of dynamic loads, be in large part of comparatively light construction, particularly in its foundation and in its general framework.

Another object of the invention and a corresponding attainment thereof, is the provision of a rock crusher and crushing process in which the rock is subjected to sonic wave trains of compression and tension, and fragmentation of the rock is attained by fatigue failure of the rock under moderate cyclic stress, rather than by application of a large crushing force, in a brute force style of action. The advantage flowing from this novel type of rock crushing include great reduction in the magnitude of required jaw force, and consequent reduction in necessary bulk and mass of the jaws, as well as throughout the entirety of the machine.

In another manner of speaking, the rock particles to be crushed are, in accordance with the broad principle of the invention, acoustically coupled into an acoustic circuit, where they are continuously subjected to an acoustic wave action which gradually reduces their size owing to progressive fragmentation by elastic fatigue failure. The rock particles remain in the acoustic circuit, and subject to continuin elastic fatigue failure, as they are progressively reduced in size, until a predetermined fineness is reached. A preferred form of crusher in accordance with the invention utilizes a wedge-shaped path or slot between the jaws for the rock passing through the acoustic circuit of the crusher. The rock contains particles originally of too great a size to more than just enter into the large upper end of this wedge-shaped slot; but as it is progressively reduced in particle size it falls gradually through the slot under the influence of gravity. It remains acoustically coupled in the circuit until it finally falls from the narrow end of the wedge-shaped slot between the jaws.

The invention will be more fully understood from the following detailed description of certain illustrative embodiments thereof, reference for this purpose being had to the accompanying drawings, in which:

FIG. 1 is a longitudinal medial section through an illustrative rock crusher in accordance with the invention, certain parts in the medial plane of the section being shown in elevation, and the near side cover of the wave generator being removed;

FIG. 2 is a plan view of the rock crusher of FIG. 1;

FIG. 3 is a section taken on line 3-3 of FIG. 1;

FIG. 4 is a broken longitudinal sectional view of another embodiment of the invention, being a section on broken line 4-4 of FIG. 5;

FIG. 5 is a plan view of the rock crusher of FIG. 4, the hopper being removed;

FIG. 6 is a section taken on line 66 of FIG. 4; and

FIG. 7 is a section taken on line 7-7 of FIG. 5.

Referring first to the embodiment of the invention shown in FIGS. 1 and 2, a relatively light base 19 is provided, comprising in this case two longitudinal earth engaging skids in the form of channels 11 connected at their ends by transverse end members 12. Mounted on one end of this base, on an I-bearn member 13 bridging channels 11, is a fixed crusher jaw or anvil 14 which affords a large inertia mass, and which is here shown in the form of a generally rectangular block. This jaw or anvil 14 has the inertia necessary to withstand or absorb a large periodic force impulse in the operation of the crusher without substantial yield or vibration.

Horizontally opposed to fixed jaw 14 is a vibratory jaw 15, also of large inertia mass, and in the general form of a rectangular block. This vibratory jaw 15 is mounted on one end of an elastic, longitudinally vibratory rod or shaft 16, preferably composed of steel for good elastic wave action without fatigue.

This opposite end of shaft 16 supports and is acoustically coupled to a sonic wave generator 20, designed to set up in shaft 16 longitudinally oriented, elastic, sonic wave action, of a nature to be described more particularly hereinafter.

Between jaw 15 and Wave generator 20, and preferably considerably nearer to the former than the latter, the shaft 16 is formed with a cylindrical mounting collar 21, which is embraced by the halves of a split stationary mounting block 22 fixed to the top flange of an l-beam 23 bridging frame members 11. Preferably, side plates or strays 23a are fastened at opopsite ends to stationary block 22 and stationary jaw 14 to steady these members.

In the operation of the crusher, jaw 15 vibrates through a very short displacement distance toward and from the opposed fixed jaw 14. It is here shown as vertically supported throughout this vibratory movement by sliding engagement with the top flange of an l-beam support 24 bridging the frame members 11.

In the illustrative embodiment of the invention, jaw 15 is squared off vertically, so as to present a working face 15a disposed in a vertical plane. The opposed side of fixed jaw 14 is channelled to form a steep sloping Working face 14a, with vertical edge margins 25 to confine the rock between the jaws. These conformations define a wedge-shaped path or slot S for the rock particles through the space between the jaws, the wider end of the wedge being at the top; and the rock material is in the acoustic circuit of the

acoustic components

14, 15, 15 and 2b of the rock crusher while in this slot.

A hopper 26 leading to the space or slot S between the jaws is mounted by means of arms 27 on jaw 14. The jaws are so spaced from one another that the upper side of the wedge-shaped slot S will just receive the largest rocks anticipated, while the gap at the lower end of the slot will pass rock of the maximum size desired in the output.

Vibration generator 29 is preferably of a type disclosed in my co-pending application entitled Vibration Generator for Resonant Loads and Sonic System Embodying Same, filed March 21, 1962, Serial No. 181,385.

oneness J3 It is therefore largely diagrammatically illustrated and only briefly described herein. With reference to FIG. 3, in addition to FIGS. 1 andZ, the generator 2% includes an intermediate body member or block 35, and two end plates 36 and 37, end plate 36 being removed to expose underlying members in FIG. 1. Block 35 has two raceway bores 38, one over the other, and each contains an inertia rotor 41?. Each such rotor 40 embodies an inertia roller 41, of somewhat less diameter than the corresponding raceway bore 38, and which is rotatably mounted on an axle 42 projecting axially from the hub portion of a spur gear 44, whose pitch circle is of substantially the same diameter as roller 41. Gear 44 meshes with an internal gear 45 formed or mounted within housing body member 35 concentrically with the corresponding raceway bore, and whose pitch circle is of substantially the same diameter as said bore.

Each rotor 40 is designed to turn in an orbital path about its raceway 38, with gear 44 in mesh with ring or internal gear 45, and with roller 41 rolling on the bearing surface afforded by the bore 38. To maintain the roller 41 in proper engagement with the raceway 38 while the generator is at rest, or coming up to speed, the axle 42. of the rotor is provided with an axial pin 46 which rides around a circular boss 47 projecting inwardly from sidewall 36 on the axis of the raceway bore 38.

Shaft is acoustically coupled to the generator 23 by being flange-connected at its end to the body member 35, as shown clearly in FIGS. 1 and 2.

The two rotors 46 are driven through a pair of rotatable and conically gyratory driveshafts 54, each of which has a universal joint coupling 55 to the corresponding spur gear. 44. The lower of the two shafts 54 is connected through a universal joint 56 to the extremity of a shaft 57 mounted coaxial with the lowermost raceway bore 38, and journalled in the walls of a suitable gear housing 60. The upper shaft 54 is similarly connected through a universal joint 61 to the extremity of a shaft 62 mounted coaxial with the upper raceway bore 38, and journalled also in suitable bearings afforded by gear housing 60.

Shafts

57 and 62 carry meshing spur gears 63 and 64, respectively, so that the shafts 54 and the rotors 49 turn in opposite directions. As here shown, the gear housing 69 is mounted on a stand 70, which also supports an electric drive motor 71 coupled to a spur gear 72 (PEG. 2) journalled in gear housing 60 and meshing with the spur gear 63.

The operation of the vibration generator is as follows: Rotation of shafts 54, which turn in opposite directions, rotates the two spur gears 44 around the internal gears 45, two shafts 54 each moving in a conical gyratory fashion. The inertia rollers 41 roll on the bearing surfaces 38, so that the rotors 40 move in orbital paths around the raceway 38. The centrifugal force developed by the rotors moving in these orbital paths is taken by pressure of the rollers 41 on the surfaces of the raceways 38. The rollers 41 turn at nearly the same rate of rotation as the gears 44, but with some slight variation or creep therebetween, which is accommodated by the rotatable mounting of the rollers 41 on the gear shafts 42. The two inertia rotors thus exert gyratory forces on the housing body 35. The rotors 49, however, are phased so that the vertical components of their motions will be always equal and opposed, while the horizontal components thereof will be in phase or in step with one another. This is accomplished in the original setting of the rotors by means of the interconnecting gearing. For example, as shown in FIG. 3, the two rotors may be set so that one is at its extreme uppermost position while the other is at its extreme lowermost position. Accordingly, the rotors move up and down with equal and opposed movements, and the vertical components of the reactive forces exerted thereby on the housing 35 are equal and opposed and cancelled within the housing. On the other hand, the gyrating rotors move horizontally in step with one another, so that the horizontal components of their reactive forces exerted against the housing are equal and in phase, and the reactive forces experienced by the housing 35 are therefore additive. The housing 35 therefore exerts an alternating force along a direction line perpendicular to the paper in FIG. 3, and in longitudinal alignment with the shaft 16 in FIGS. 1 and 2.

It will be observed that the preferred type of generator disclosed has a desirable frequency step-up characteristic from drive motor input to vibratory housing output force, in that for each orbital trip of a given gear 44 and its corresponding inertia roller 41 around the inside of in ternal gear and raceway bore 38, the shaft 54, gear 44 and roller 41 make only a small fraction of a complete revolution on their own axes. The shafts 54 thus gyrate in their conical paths at greater frequency than their own rotational frequency on their own axes. Thus the orbital frequency of the inertia rotors 41, and the vibration output frequency of the generator housing, is correspondingly multiplied over the rotational frequency of the drive motor. A simple low speed drive motor may thus be used, and a desirably high vibration output frequency obtained therefrom. The output frequency may be set in the design of the generator, the step up in frequency being determined by the relative diameters of the gears 44- and 45. The output frequency is at some selected value in the typical range of to 500 c.p.s., and it will be evident that, for a motor of any given speed rating, the gear ratio from motor to generator, and the step-up of frequency within the generator, may readily be made such as to furnish the desired output frequency. The frequency range quoted is typical for many common types of input and output. Some metallurgical and chemical processes desire a very fine powder. For the latter I may use known acoustic sources giving tens or hundreds of thousands of cycles per second.

From the foregoing description of the vibration generator 20, it will be understood that the effect of the operation of the latter is to apply to the extremity of elastic shaft 16 an alternating force directed along the longitudinal axis of said shaft. The shaft 16, preferably and in the illustrative embodiment, has its intermediate mounting collar 21 located substantially closer to its end coupled to the jaw 15 than to its end coupled to the generator 20. As shown in the present drawings, the mounting point is located at about 25% of the length of the shaft from the coupling point to the jaw 15, though it may be substantially closer. The generator 20 is driven to furnish an output frequency such as will set up in the shaft 16 a longitudinally oriented resonant standing wave, with a node N at the mounting collar 21 of the shaft, an antinode V at the coupling point to the jaw 15, and an antinode V at the generator housing. The prime mover 71, gearing leading therefrom, the gear ratio of generator 71, and the length and mounting point of shaft 16 are designed in relation to one another to produce the desired resonant standing wave in shaft 16, utilizing principles which are familiar to those skilled in the art. This standing wave is in general of half-wavelength character, in that it has velocity antinodes at its ends and an intervening node. The wave pattern is modified, however, by location of the fixed mounting point for the shaft 16 suificiently closer to one end of the shaft than the other, and by the large mass of the jaw 15, so that its actual length is closer to one-quarter-wavelength. The standing wave pattern obtained is diagrammed in FIG. 1, just above the shaft 16, the vertical height of the pattern at any point along its length being representative of both the amplitude and velocity of longitudinal vibration at the corresponding point of the acoustic system or circuit comprised of the shaft 16, jaw 15 and generator 21 As will be understood, and as is evident from the standing wave pattern diagrarnmed in FIG. 1, the amplitude of longitudinal vibration at the mounting point of the shaft 16 is substantially zero, affording the aforementioned node N. The two arms 16a and 16b of the elastic shaft 16 elastically elongate and contract in unison with one another in the establishment of the standing wave pattern, the extremities of the arms 16a and 161) having relative amplitudes as represented by the standing wave diagram above the shaft. As will further be evident, the amplitude of the vibratory motion is considerably larger at the generator end of the shaft 16 than it is at the jaw 15. Correspondingly, the cyclic force exerted by the shaft arm 160 on the jaw 15, and in turn by the jaw on the rock in the slot S, is proportionately multiplied over the cyclic force exerted by the generator 2h on the generator end of the shaft. At the same time, the velocity, or displacement amplitude, of the large, inertia mass jaw 15 is relatively low. The rock wedged between the

inertia mass jaws

14 and 15 is thus subjected to a cyclic stress of high magnitude, but with displacement amplitude and velocity at a very low magnitude. The condition at both the jaw 15 and within the rock between the jaws is thus one of high acoustic impedance. Under these conditions, the rock undergoes an alternating compressional and tensional wave, with the magnitude of cyclic tension materially exceeding the endurance limit of the rock, so that the rock fails quickly by elastic fatigue, and shatters rapidly into smaller and smaller particles. The rock is indicated generally at R in FIG. 1, and will be seen to enter the slot S from hopper 26 in relatively large particles, which are progressively reduced in size by the action of the jaws, and fall by gravity from the lower end of the slot S, reduced to a predetermined maxirnum size, as will be clear from PEG. 1. The rocx particles will be seen to be in the acoustic circuit of the rock crusher during this progressive reduction in size owing to the wedge shape of the slot. The described high acoustic impedance at the movable jaw 15 is desirable for good impedance match to the rock. The desirable high impedance is attained by using a jaw 15 of large inertia mass, and therefore high mass reactance. By providing for a mass reactance which is large as compared with the resistive vector component of the impedance (which resistive component is of course owing to frictionfl dissipation of energy in the process) the Q factor of the vibratory system is desirably high. The factor Q will of course be understood to be a figure of merit of vibratory systems, measured either by the ratio of the reactive component of impedance to the resistive component thereof, or by the ratio of energy stored to energy expended per cycle of operation. An additional advantage of the provision of a high impedance at the movable jaw and within the rock, and a considerably lower impedance at the vibration generator, is that the generator can then operate easily with high mobility, under practical conditions of lower force and higher velocity that is requisite at the crusher jaw. -t can also be driven readily from simple and conventional prime movers. The system is thus characterized by desirably low impedance at the generator end, and desirably high impedance at the rock crushing end, with the intervening elmtically vibratory shaft 16 functioning as an acoustic lever, or in another concept, as m impedance adjusting transformer.

While the node N for the shaft 16 is here shown as located approximately of the length of the shaft from the inertia-mass jaw 15, it can, in practice, he considerably closer, with desirably further increased output impedance. The jaw 15 may thus have its amplitude of vibration reduced to a very small magnitude. The total length of the standing wave is then quite close to a quarter-wavelength, and from a practical standpoint, the standing wave may be said to be approximately a quarterwavelength long. Even in such case, however, the standing wave system comprises two velocity antinodes and an intervening node, so that, while the actual distance from antinode V to node N may become quite small, the standing wave is in the nature of a half-wave system in the sense that it has opposed motion at its ends, and an intervening node. And, of course, harmonic frequency standing waves are quite possible, and comprise modifications within the scope of the invention.

In FIG. 1 there is illustrated a water discharge pipe 69, leading in to hopper 26. In one practice of the inven tion, water can thus be run through the slot 8 during the treatment, aiding in cleaning the rock of dirt and organic material, and in moving the smaller material downward through the slot. This water also acts as a coupling medium between the face of the jaw 15 and the rock. Without the water, necessary acoustic coupling arises from the rock becoming wedged between the two jaws. With Water present, sonic waves are radiated from the movable jaw into the water, and thence to rock particles not in direct contact with the jaws. The sonic waves then traverse such rock, and subject it to a cyclic stress leading to fatigue failure. Also, the sonic waves in the water surrounding the rock have a sonic cleansing action on the rock, removing dirt and organic material from the rock, and washing the same down and out of the crusher.

In FIGS. 47, I have shown a modification in which the stationary inertia-mass jaw of FIGS. 1 and 2 is replaced by a second vibratory jaw, connected into a dual acoustic circuit so as to vibrate in opposite phase to the first vibratory jaw.

A skid type base 80 is provided, generally similar to the base 10 of the first described embodiment. Near one end of this base is a vibratory, inertia mass jaw 81. This jaw 81 is mounted on one end of an elastic, longitudinally vibratory shaft 82, the opposite end of which supports and is acoustically coupled to a sonic wave generator $3, which may be precisely similar to the generator it? of FIGS. 1 to 3. In the case of the present embodiment, however, the generator 83 is oriented in a plane at right angles to that of the generator 29 of FIGS. 1 to 3. Thus, the two rotors 84 of generator 83 are on horizontally or laterally spaced vertical axes, as will be clear from an inspection of FIG. 5. It will further be clear then, notwithstanding the turning of the generator through 90 degrees, the generator 83 operates, in the general manner of generator 2% of FIGS. 1 to 3, to apply to the end of shaft 82 an alternating force directed longitudinally of said shaft. Between jBIW 81 and generator 33 the shaft has cylindrical mounting collar 35, located, as in the case of FIGS. 1 and 2, substantially closer to jaw 81 than to generator 83. This mounting collar 85 is embraced by split mounting block 86 fixed to the top flange of I-beam 37 supported on frame 80.

Horizontally opposed to vibratory inertia mass jaw 81 is a vibratory inertia mass jaw 83, in the general form of a rectangular block, as clearly appears in FIGS. 4 and 5. This vibratory jaw 88 is mounted on the ends of a pair of elastic, longitudinally vibratory shafts 89, extending generally parallel to and on opposite sides of shaft 82, outside or beyond the two ends of the vibratory jaw 81, as clearly appears in FIG. 5. The opposite end of each of shafts 89 is acoustically coupled to and supports a sonic wave generator 99, preferably constructed like generators 29 and 83 and oriented like the generator 83. Each of the two shafts 39 is of approximately half the cross-sectional area of the shaft 82, and each generator tl is dimensioned to generate approximately half the output force of the generator 33. The shafts 89 are provided with mounting collars 91 embraced by split mounting blocks 92 fixed to the top flange of an I-beam 93 mounted on base 8% the collars be ng located closer to jaw 88 than to generators 99, in substantially the same proportionate location as that of collar on shaft 82.

The inertia jaw 88 is dimensioned to have approximately the same mass as inertia jaw 81. In the embodiment of FIGS. 4 and 5, jaw 88 is squared off vertically, so as to present a working face 88a in a vertical plane. The opposed side of jaw 81 is channelled to form steep sloping working face 81a, with vertical edge margins 94 to confine the rock between the jaws. These confirmations define a wedge-shaped path or slot S for the rock particles through the space between the jaws, the wider end of the wedge being at the top. As in the earlier embodiment, the rock material is in the acoustic circuit of the rock crusher while in this slot. A hopper 95 leading into slot S is supported by standard 96 erected from the base 81}. The jaw faces 81a and 88a are spaced from one another at the top to receive the largest rock particles anticipated, while at the bottom the spacing is such as to pass material ground or crushed to the fineness desired.

An internal combustion engine 190, serving as a prime mover, is mounted on base I-beams 101, and the driveshaft of this engine carries bevel gear 104 in mesh with bevel gear 105 on a vertical shaft 166 (FIG. 7) journalled suitably in the upper and lower walls of a gear housing It)? mounted on one end of base 80. The gear housing Hi7 includes a removable cover 163, some of which has been broken away in FIG. 6.

Bevel gear shaft 106 carries, within housing 107, a spur gear 119, which meshes with a spur gear 111 on a shaft 112, and this shaft 112 is coupled through universal joint 113, conically gyratory driveshaft 114 and a universal joint 115 to one of the rotor shafts of generator 83. Generator 83, like generator 29, has two inertia rotors, such as $4, each driven by a driveshaft such as the shaft 114 shown in FIG. 7. A spur gear 116 meshing with spur gear 111, and therefore turning in the opposite direction, will be understood to drive the second inertia rotor 84 of the generator $3 through a second conically gyratory driveshaft and universal joint coupling arrangement, not shown, but identical to that illustrated in FIG. 7. The inertia rotors 84 will be understood to be phased like those of the generator of FIGS. 1-3, so as to produce an alternating output force applied to the end of shaft 82.

The gear 119 on bevel gear shaft 196 also meshes with a gear 120 on a housing-journalled shaft 121 that is coupled through universal joint 122, conically gyratory driveshaft 123 and universal joint 124 to one of the inertia motor 125 of one of the sonic wave generators 9d (the one towards the bottom of the drawing as viewed in FIGS. and 6). Fast with gear 129 is a smaller gear 126, meshing with an equal sized gear 127 on a housingjournalled shaft 128 (FTG. 6), and this shaft 128 will be understood to drive the second inertia rotor 125 of generator 90 through a second conically gyratory drive shaft and universal joint coupling arrangement, not shown, but understood to be identical to that shown in FIG. 7.

Thegear 116 also drives a gear train operating the second sonic wave generator hi), all in a manner substantially identical to the case of the first generator 90. Thus, as appears in FIG. 5, gear 116 meshes with a gear 110,

' which is identical to gear 110 excepting that it is an idler rather than a power input gear; and gear 110' drives gear 129', with which turns a gear 126 meshing with and driving a gear 127'. The gears 126' and 127 drive the inertia rotors 125 of the second generator 90 through shafts, universal joints, and a conically gyratory shaft arrangement such as seen in FIG. 7, and which need not be further illustrated.

The pair of inertia rotors 125 of each of sonic wave generators 90 are phased for cooperation to generate an alternating force longitudinally of the corresponding shaft 89. That is to say, the rotors 125 of each generator are interconnected with one another through their gearing arrangements so as to move longitudinally of shaft 89 in unison with one another, and transversely of shaft 89 in opposition to one another. Thus the force components longitudinal of the shaft 89 are additive, while those transversely of the shaft are cancelled. Moreover, the inertia rotors 125 of both generators 99 are interconnected by the intervening gear train to be in phase with one another, so that the alternating forces applied by the two wave generators to the shafts 89 will be in phase; while the inertia rotors 84 of the sonic wave generator 83 are interconnected in the gear train to be in 180 phase opposition to the inertia rotors of the generators N. Accordingly, the alternating force applied to the end of shaft 82 is in 180 phase opposition to the alternating forces applied to the corresponding ends of the two shafts 89. Therefore, the inertia masses 8]. and 88, which comprise the two jaws of the rock crusher, vibrate in opposed phase, or in opposition to one another.

From the foregoing, it will be clear that the acoustic system comprised of the inertia mass jaw 81, the elastic shaft 82;, stationarily mounted at 85, 86, and the sonic wave generator 83, when driven by generator 83 at a resonant standing wave frequency for the system, vibrates in a half-wavelength type of standing wave pattern such as represented in FlG. 1, there being a velocity antinode at V, a node at N, and a velocity antinode at V'. In a similar manner, the acoustic vibratory system comprised of the inertia mass jaw 88, the two elastic shafts 89, stationarily mounted at 91, 2, and the two sonic wave generators 9%, provide a half-wavelength type of standing vave pattern, also like that diagrammed in FIG. 1, when the generators 9d are operated at the resonant standing wave frequency of this system. An inspection of FIG. 5 will reveal that the masses and dimensions of the two systems, i.e., that comprised of jaw 81, shaft 82, and generator 83, and that comprised of jaw 88, the two shafts 89, and the two generators 9d, are sufficiently comparable that a good resonant standing wave pattern can boot)- tained in both systems in a common properly chosen frequency range. The previously described gear train interconnecting the several generators will be seen, from FIGS. 5 and 6, to besuch as to drive the several genera- .tors at a common frequenc and internal combustion engine 1% will be understood to be operated at a controlled speed to accomplish resonant standing wave behavior in both systems. Since the alternating forces applied to the shaft 82 and to the shafts 89 are 180 out of phase, the two

inertia mass jaws

81 and 88 vibrate towards and from one another in 180 opposition. Rock material in the wedge-shaped slot S between the jaws is accordingly subjected to the high forces incident to the low amplitude vibratory movements of these inertia mass jaws. The detailed discussions given in the introductory portion of this specification and in connection with the first described embodiment (FIGS. 1-3) as to conditions of acoustic impedance, extremely high force application but low amplitude of vibratory movement at the jaws, wave pattern and wave length characteristics, etc, apply equally here. The essential difference is simply, that, in the present case, both jaws vibrate, instead of one being vibratory and the other stationary. In the first system, in effect, the rock material is between a highinertia anvil member and a high impedance vibratory output member of an elastically vibratory sonic system. In the second instance, the rock material is between two high impedance vibratory output members of two elastically vibratory sonic systems operating in 180 phase opposition.

In the foregoing description, it will be evident that both of the systems described accomplish the objectives preliminarily stated, and embody the various features of advantage described in the int oductory portion of the specification. Rock material is crushed rapidly and dimin ished in size to predetermined fineness using machinery which can be greatly lightened in many of its parts as compared with conventional rock crushers, and which can easily be made portable for ready transportation to or about the site of operations.

It will be understood that the drawings and description are merely for illustrative purposes, and that various changes in design, structure and arrangement may be made without departing from the spirit and scope of the invention as defined by the appended claims.

I claim:

1. In a crushing apparatus of the character described:

a pair of opposed, massive crusher jaws between which a substance to be crushed may be positioned, and at least one of which is vibratory toward and from the other,

a sonic wave generator adapted to deliver an alternating output force, and

an elastically vibratory wave transmission system comprising a member of solid elastic material having a range of elastic vibrations intercoupled between said generator and said vibratory jaw, so as to receive said alternating force, undergo corresponding elastic vibration, and impart vibration to said vibratory jaw 2. The subject matter of claim 1, wherein said jaw! are horizontally opposed to one another,

said vibratory jaw is movable by said wave transmission system with a horizontal component of vibration, and

said jaws present to one another opposed faces which define a downwardly directed wedge-shaped passage for the substance to be crushed.

3. The subject matter of claim 1, wherein said elastically vibratory transmission system is characterized by relatively low acoustic impedance where coupled to said sonic wave generator, and relatively high impedance where coupled to said vibratory jaw, whereby relatively high vibration amplitude and low cyclic force at the generator is transformed into relatively low vibration amplitude and high cyclic force at the vibratory jaw.

4. The subject matter of claim 1, wherein said elastically vibratory transmission system comprises an elongated longitudinally elastically vibratory structure coupled at one end to said sonic wave generator, and at its other end to said vibratory jaw, and

means operating said sonic wave generator at a resonant longitudinal standing wave frequency for the acoustic system comprised of said vibratory jaw, elongated vibratory structure, and sonic wave generator.

5. The subject matter of claim 1, wherein said elastically vibratory transmission system comprises an elongated longitudinally elastically vibratory structure coupled at one end to said sonic wave generator, and at its other end to said vibratory jaw,

and including means affording a stationary nodal point support for said elongated structure at a point substantially nearer its end coupled to said vibratory jaw than to its end coupled to said sonic wave generator, and

6. The subject matter of claim 5, wherein the second of the massive opposed crusher jaws is mounted to function as an anvil.

7. The subject matter of claim 1, wherein said jaws are horizontally opposed to one another, with a space for material to be crushed therebetween, and said vibratory jaw is movable with a horizontal component of vibration,

said elastically vibratory transmission system comprising a generally horizontally elongated longitudinally vibratory structure coupled at one end to said sonic wave generator, and at its other end to said vibratory j means affording a stationary nodal point support for said elongated structure at a point substantially nearer its end coupled to said vibratory jaw than to its end coupled to said sonic wave generator, and

1% means operatin said sonic wave generator at a resonant longitudmaly standing wave frequency for the acoustic system comprised of said vibratory jaw, elongated vibratory structure, and sonic wave generator.

8. The subject matter of claim 7, wherein the second of the massive crusher jaws is mounted to function as an anvil.

9. The subject matter of claim 1, wherein said crusher jaws are both vibratory toward and from one another, and there is a sonic wave generator for each jaw, and an elastically vibratory wave transmission system intercoupled between each generator and the corresponding vibratory jaw, and

means interconnecting the sonic wave generators in a phase relationship producing opposed vibratory movements of the jaws toward and from one another.

10. The subject matter of claim 9, wherein:

said jaws are horizontally opposed to one another,

said vibratory jaws are movable by said wave transmission systems with horizontal components of vibration, and

11. The subject matter of claim 9, wherein said elastically vibratory wave transmission systems have relatively low acoustic impedance where coupled to said sonic wave generators and relatively high impedance where coupled to said jaws.

12. The subject matter of claim 9, wherein,

said elastically vibratory wave transmission systems comprise elongated longitudinally elastically vibratory structures, each coupled at one end to the corresponding wave generator and at the other to the corresponding jaw,

the two acoustic systems comprised of the two elastically vibratory wave transmission systems taken together, in each case, with their corresponding wave generator and jaw, being resonant in a common frequency range, and

means operating said wave generators in said common frequency range.

13. The subject matter of claim 12, including:

means aiford ng a stationary nodal point support for each of said elongated longitudinally elastically viratory structures at a point substantially nearer its end coupled to the corresponding jaw than to its end coupled to the corresponding wave generator.

References fired in the file of this patent UNITED STATES PATENTS 962,998 Christ et al June 28, 1910 2,198,148 Bailey Apr. 23, 1940 2,258,059 Kessler Oct. 7, 1941 2,960,314 Bodine Nov. 15, 1960 OTHER REFERENCES Russia, 122,663, application date Dec. 15, 1956, approved for printing Aug. 7, 1959.

Claims

1. IN A CRUSHING APPARATUS OF THE CHARACTER DESCRIBED: A PAIR OF OPPOSED, MASSIVE CRUSHER JAWS BETWEEN WHICH A SUBSTANCE TO BE CRUSHED MAY BE POSITIONED, AND AT LEAST ONE OF WHICH IS VIBRATORY TOWARD AND FROM THE OTHER, A SONIC WAVE GENERATOR ADAPTED TO DELIVER AN ALTERNATING OUTPUT FORCE, AND AN ELASTICALLY VIBRATORY WAVE TRANSMISSION SYSTEM COMPRISING A MEMBER OF SOLID ELASTIC MATERIAL HAVING A RANGE OF ELASTIC VIBRATIONS INTERCOUPLED BETWEEN SAID GENERATOR AND SAID VIBRATORY JAW, SO AS TO RECEIVE SAID