Hydroacoustic apparatus

Abstract

A class of hydroacoustic oscillators is described having a fluid pressure-actuated valving mechanism which provides self-excited fluid pressure variations on an oscillator mass. The pressure actuated valving mechanism switches pressure abruptly, first from low to high values to obtain driving forces for accelerating the oscillator mass from a displaced position into a spring system, and second from high to low values to enable the spring to decelerate the oscillator mass to zero velocity, and then to return the mass with increasing acceleration in the opposite direction towards its displaced position. Motion of the oscillator mass is coupled to the valving mechanism, without restricting the displacement of the oscillator mass, to cause self-excited fluid pressure variations to enable the oscillation of the mass to be sustained.

Description

The present invention relates to improved hydroacoustic apparatus and particularly to an improved class of hydroacoustic ocsillators having pressure actuated valving mechanisms which switch driving pressures to excite the oscillators.

Hydroacoustic apparatus provided by the invention are suitable for driving external loads, for the radiation of acoustic energy, for vibration testing and especially for delivering a rapid succession of high energy impacts to an anvil system as may include a shank, drill steel and bit of a percussive drill or the pile cap of a pile or the moil of a demolition tool.

It is an object of the invention to provide improved hydroacoustic oscillators.

It is another object of the invention to provide improved hydroacoustic oscillators having improved hydraulic pressure actuated valving mechanisms.

It is a further object of the present invention to provide improved hydroacoustic oscillators having improved valving structures which switch driving pressures in an abrupt manner to self-excite the oscillators.

It is a still further object of the present invention to provide an improved hydroacoustic oscillator having an improved valving structure which fully opens and fully closes inlet and outlet ports to the oscillator during each cycle of oscillation and which opens and closes such ports during intervals which are short as compared to the period of oscillation.

It is a still further object of the present invention to provide an improved hydroacoustic oscillator having high efficiency of conversion of hydraulic energy to oscillatory mechanical energy.

It is a still further object of the present invention to provide an improved hydroacoustic oscillator wherein power loss in porting of the inlet and outlet flow of the oscillator is minimized.

It is a still further object of the present invention to provide an improved hydroacoustic oscillator valving mechanism which switches driving pressures within the oscillator when the oscillating mass achieves its maximum velocity.

It is a still furhter object of the present invention to provide an improved hydroacoustic oscillator having a controlled percussive force generation cycle.

It is a still further object of the present invention to provide an improved hydroacoustic oscillator valving mechanism which switches driving pressures to the oscillator so as to obtain approximately square wave driving forces.

It is a still further object of the present invention to provide an improved hydroacoustic oscillator which has a relaxation mode of oscillation for providing an advantageous relationship of energy output to oscillator size.

It is a still further object of the present invention to provide an improved hydroacoustic oscillator having high efficiency of power conversion in the presence of asymmetrical oscillation waveforms as may occur with non-linear or time varying loads such as may be presented when the oscillator mass experiences impact events.

Briefly described, a hydroacoustic oscillator embodying the invention includes an oscillator mass, a spring member, and a pressure-actuated valving mechanism coupled to the mass in a manner for controlling the timing of valving operations. The pressure actuated valving mechanism controls the flow of pressurized fluid from a region of relatively high pressure, through the oscillator, to a region of relatively low pressure, thereby to produce periodic oscillations of the mass at a frequency determined by the oscillator mass and the characteristics of the spring member communicating therewith. The valving structure includes a valving element actuated by the oscillator mass to provide a mechanism which ports, sequentially, and in predetermined time relationship with the mass displacement, first the inlet flow at high pressure to one side of the mass, and secondly the discharge flow at low pressure away from said one side of the mass so as to provide fluid pressure variations on the mass that alternate between relatively high and relatively low pressure, thereby to actuate the mass. The resulting pressure variations approximate a square wave driving force on the mass resulting from the switching of the ports from the high and low pressure regions. To obtain maximum efficiency of power conversion the valving element fully opens and fully closes the inlet and outlet ports, respectively, in time intervals that are short compared to the period of oscillation. The valving element may also provide the port switching action during intervals when the mass has achieved predetermined velocities during its cycle of oscillation, as for example, when the mass has achieved its maximum velocity in either direction.

The approximately square wave actuating force generated by the port switching action of the valving element is fed back upon the oscillator mass-spring system, which system may be provided by the oscillating mass and a pressurized fluid-filled cavity to which one end of the mass is exposed, so as to sustain self-excited oscillatory motion of the mass-spring system at or near a defined oscillation frequency.

The pressurized fluid-filled cavity and the valving structure are on opposite sides of the mass. The cavity provides a bias or restoring force on one side of the mass whose average value balances the average value of the square wave driving force applied to the opposite side of the mass, due to the valving action, and the average of any externally applied force, so that there is thereby maintained a desired equilibrium or average position of the mass.

In one embodiment of the invention the dynamic spring rate or stiffness of the pressurized spring portion of the system is reduced toward zero so as to provide a constant spring force upon the oscillator mass, independent of the instantaneous position of the mass over a cycle of oscillation. In this instance, the oscillator behaves as a relaxation oscillator, and its frequency is dependent upon the supply pressure. The relaxation mode of oscillation is especially useful with certain non-linear or time-dependent loads, such as may be produced when the oscillator mass experiences impact events. This embodiment has as a feature the achievement of maximum energy transfer to the load within given miminum oscillator dimensions (viz., the space occupied by the oscillator).

The foregoing other and additional objects, advantages and features of the invention will become more readily apparent from a reading of the following specification in connection with the accompanying drawings in which:

FIG. 1 is a transverse sectional view of a hydroacoustic oscillator embodying the invention;

FIG. 2 is a plan view of the valving element utilized in the hydroacoustic oscillator shown in FIG. 1;

FIGS. 3 and 4 are fragmentary sectional views of the hydroacoustic oscillator shown in FIG. 1 at positions of maximum displacement during the cycle of oscillation;

FIG. 5 is a series of waveforms illustrating the variations in displacement pressure and forces resulting during the operation of the hydroacoustic oscillator shown in FIGS. 1 through 4;

FIG. 6 is a transverse sectional view of a hydroacoustic oscillator provided in accordance with another embodiment of the invention and also illustrating the oscillator included in an impact tool;

FIGS. 7A, 7B are respectively plan and cross-sectional views of the valving element used in the oscillator shown in FIG. 6;

FIGS. 8, 9 and 10 are fragmentary sectional views showing the oscillator illustrated in FIG. 6 in various positions during its cycle of oscillation;

FIG. 11 is a series of waveforms illustrating variations in displacement, driving forces and pressures resulting from the operation of the oscillator illustrated in FIGS. 6 through 10;

FIG. 12 is a transverse sectional view of a hydroacoustic oscillator in accordance with still another embodiment of the invention;

FIG. 13 is a transverse sectional view of a hydroacoustic oscillator in accordance with a still further embodiment of the invention; and

FIG. 14 illustrates a cavity structure which may be used in the hydroacoustic oscillator shown in FIGS. 12 and 13.

Referring now to the drawings, the hydroacoustic oscillator assembly shown in FIG. 1 has a housing 10 of generally cylindrical shape. Bolts, screw threads, and other fastening devices which are used in the construction are not shown to simplify the illustration. The housing has a central bore 12 in which a driven mass 14 and cooperating valving element 16 are free to oscillate. The driven mass 14 and the valving element 16 are supported in their motions on a thin lubricating fluid film between their cylindrical outer walls and the wall of the bore 12. This lubricating film is provided from the flow of pressurized fluid which drives the oscillator. This fluid is preferably hydraulic oil. The mass 14 is a piston having a main or larger diameter portion 18 which fits in the diameter of the bore 12. The mass 14 is also referred to herein as the piston 14. The piston 14 also has a smaller diameter portion which extends downwardly through the bottom of the housing 10 and thus provides a shaft 20 for delivering output power to a load. Suitable seals and/or packing may be provided in the region of the housing through which the shaft penetrates so as to prevent the escape of fluid from the bore. The valving element 16 is engaged by two rings 22 and 24 which are fixedly attached to the shaft 20. These rings serve to position the valving element 16 with respect to inlet and outlet ports 26 and 28 which cooperate with the valving element 16 to provide a valve mechanism or structure for the oscillator.

The ports 26 and 28 are circular grooves which extend through the wall of the bore 12. The inlet port 26 communicates with fluid pressure supply means while the outlet port 28 communicates with discharge or return means for the fluid pressure. The valving element 16 is in the form of a circular sleeve, shown in top view in FIG. 2. The interior portion of the sleeve is relieved by means of a plurality of semicircular, longitudinal grooves 30. These grooves and the periphery of the shaft 20 provide a plurality of passages or channels which allow fluid on either side of the valve element 16 to circulate back and forth freely, while cusps between the grooves 30 provide interference for engagement with the rings 22 and 24.

The housing has a first cavity 32 which is exposed to the upper end of the piston 14. The cavity 32 is also filled with fluid (hydraulic oil), and provides the pressurized fluid spring portion of the oscillation system. The mass of the piston 14 and the stiffness of the fluid in the cavity 32 define the natural or resonant frequency of the oscillator. This frequency may be designated f.sub.o where ##EQU1## where .rho. c equals the bulk modulus of the fluid in the cavity 32, M.sub.P equals the mass of the piston, A.sub.P equals the area of the piston exposed to the fluid in the cavity, and V equals the volume of the cavity 32. A.sub.P is defined by the expression ##EQU2## where D.sub.P is the larger diameter of the piston 14.

Another or second cavity 34 is defined in the housing in the portion of the bore 12 surrounding the shaft 20. It is through this second cavity that fluid is caused to flow under the control of the valving mechanism which actuates the oscillator.

The end of the piston 14 exposed to the cavity 34 contains only the differential area 17 between the area of the piston and the area of the shaft. This differential area 17, being exposed to the cavity 34, acts in many respects as the end of the piston, and is for convenience referred to herein and in the appended claims as an end of the piston, and is of course the end opposite to that end of the piston 14 which is exposed to the cavity 32.

An inlet line 36 feeds a supply cavity 38 with pressurized fluid, as from a hydraulic pump. This cavity 38 is of a size so as to provide an energy storage reservoir function. The line 36 supplies the steady state flow demand of the oscillator. The instantaneous, peak flow demands may be supplied from the cavity 38. The cavity 38 communicates with a cylindrical channel 40 which narrows into the groove providing the inlet port 26. The supply cavity 38 is thus closely coupled to the inlet port 26.

Similarly closely coupled to the outlet or discharge port 28 is a discharge or return cavity 42. The discharge cavity 42 communicates with a cylindrical channel 46 which narrows into the groove 28 defining the outlet port 28. This cavity is connected to an outlet or discharge line 44 which may be connected to a tank or reservoir or to the return side of the pump which feeds the inlet line 36. The cavity 42 may be located on the opposite side of the housing from the supply cavity 38. The cavity 42 is of a size to function as an energy storage reservoir which accepts peak discharge demands of the oscillator while providing a steady state discharge flow through the outlet line 44. Thus the flow into the supply cavity 38 will sustain an approximately steady supply pressure therein while the flow out of the oscillator into the discharge cavity 42 is met by an approximately steady return pressure. The supply pressure may be expressed symbolically as P.sub.S while the return pressure may be expressed by P.sub.R. These pressures may be adjusted or predetermined so as to control the output power desired from the oscillator.

The cavity 32 is fluid filled by a line 48 which is quite narrow and may be further controllably restricted by a control valve 50. This line 48 communicates the supply cavity 38 with the cavity 32. The cavity 32 also communicates with the discharge cavity 42 by way of a line 52. The line 52 is also restricted by a narrow passageway or restrictor 54 therein. The lines 48 and 52 provide a pressure divider network which is adjusted using the valve 50 to control the equilibrium position of the piston 14 by setting the average (downward) force exerted on the piston by the average pressure in the cavity 32 to equal the average (upward) force exerted on the differential area of the piston exposed to the cavity 34 together with any external average thrust exerted upwardly on the shaft 20, as by the load driven by the shaft.

The differential area, A.sub.D, of the piston 14 exposed to the pressures in cavity 34 is ##EQU3## where D.sub.P is the diameter of the piston 14 and D.sub.S is the diameter of the shaft 20. In the absence of external thrust on the shaft 20, the force balanced condition is

P.sub.C A.sub.P = P.sub.D .sup.. A.sub.D (4)

where P.sub.C is the average pressure in the cavity 32 and P.sub.D the average pressure in the cavity 34. The control valve 50 is set so that equation (4) is satisfied and the piston 14 is maintained in static equilibrium.

The axial length of the valve element 16, is the embodiment of the invention illustrated in FIG. 1, is equal to the distance separating the outer or furthest apart metering edges 27 and 29 of the inlet and outlet ports 26 and 28 respectively. In the equilibrium position as shown in fig. 1, the upper and lower edges edges. the sleeve, which provide the valve element 16, are in line-to-line relationship with the inlet and outlet port metering edtes. This structural relationship provides a high pressure gain condition in that a small axial displacement of the piston in either direction will result in a substantial change of pressure in the cavity 34, which will then act to vary the force applied to the piston 14.

The operation of the oscillator illustrated in FIG. 1 may be better understood with reference to FIGS. 3 and 4 which illustrate the piston in its maximum upper and lower positions, respectively.

Consider the case where the piston 14 has an initial displacement in a downward direction from its equilibrium position as shown in FIG. 1. The corresponding downward motion of the valve element 16 opens the inlet port 26 so that full supply pressure which is present in the supply cavity 38 is applied to the cavity 34. The pressure in the cavity 34 then rises toward supply pressure P.sub.S. The increased pressure acts upon the differential area A.sub.D and causes the piston 14 to accelerate in the upward direction. As the piston 14 passes the equilibrium position, the inlet port 26 closes and the outlet port 28 opens. The pressure in the cavity 34 then falls toward return pressure. The piston 14 continues travelling in the upward direction, but decelerates due to the reduced pressure in the cavity 34 until the piston stops. The forces due to the fluid spring in the cavity 32 are fed back and cause the piston to return in a downward direction toward the equilibrium position. As the piston passes the equilibrium position the inlet port 26 opens and the outlet port 28 closes. The driving pressure applied to the cavity 34 then increases towards P.sub.S and the downward motion of the piston 14 is decelerated in response to the driving pressure.

The piston 14 does not reach the maximum position shown in FIGS. 3 and 4 initially. Rather the amplitude of oscillation builds up gradually over an initial number of cycles.

When unloaded the oscillation frequency will be substantially the resonant frequency f.sub.o defined by equation (1). This frequency may be subject to some modification due to the acoustic impedance presented by the ports 26 and 28 and the fluid in the supply and return cavities 38 and 42. As long as the impedances presented by the ports are small as compared to the impedance presented by the piston 14, the oscillation frequency is a close approximation to f.sub.o as defined by equation (1). While the maximum amplitudes of the oscillation (viz., the positions shown in FIGS. 3 and 4) are determined by system non-linearities, it may be desirable to provide stops for limiting the upward and downward movement of the piston beyond the positions shown in FIGS. 3 and 4.

During oscillation, the piston 14 executes simple harmonic motion. The waveform of this motion is symmetrical about the mean position, which is the equilibrium position shown in FIG. 1. The maximum velocity of the piston 14 occurs as the piston, when travelling in either direction, passes through the equilibrium position. Opening and closing of the ports 26 and 28 (port switching) thus occurs when the piston is moving at maximum velocity and therefor in the shortest time period possible. This feature of operation of the oscillator provided by the invention enables the oscillator to attain high power conversion efficiency.

Waveforms (a) to (f) of FIG. 5 are further explanatory of the operation of the oscillator illustrated in FIGS. 1 through 4. Waveform (a) shows the displacement X.sub.P of the piston. The positive direction of X.sub.P is taken with the piston moving upwardly into the cavity 32. It will be noted from waveform (a) that the maximum piston velocity occurs when the piston passes through the midpoint or equilibrium position (X.sub.P = 0).

Waveform (b) shows the time history of the force F.sub.SP on the piston due to pressure variations in the upper cavity 32. These pressure variations result from the motion of the piston 14 and from the steady state pressure or pressure bias due to the divider network consisting of the channels 48, the valve 50, the channel 52 and the restrictor 54. F.sub.SP can be expressed by the following equation:

F.sub.SP = P.sub.c .sup.. A.sub.P (1 + .alpha. sin .omega.t) (5)

where .alpha. is a modulation coefficient depending upon the particular dimensions of the oscillator and the bulk modulus of the fluid. The average or steady pressure bias due to the pressure divider is indicated in waveform (b) as F.sub.SP. It is the instantaneous force F.sub.SP which is fed back during oscillation and which helps to produce self-sustained or free-running oscillation in the system.

The counterbalancing force to F.sub.SP is the force F.sub.D on the differential area A.sub.D exposed to the cavity 34. This force F.sub.D is illustrated in waveform (c). The force F.sub.D has two states, namely, F.sub.D = P.sub.S .sup.. A.sub.D or F.sub.D = P.sub.R .sup.. A.sub.D, since the cavity is either open to supply pressure or to return pressure. Assuming for simplicity that the return pressure P.sub.R = 0, then the variation of F.sub.D with time, as illustrated in waveform (c), exhibits abrupt variations in switching from zero to full driving force.

The net force F.sub.N driving the piston 14 is the difference between F.sub.D and F.sub.SP. This force is plotted in waveform (d). The average value of F.sub.N with no external forces (viz., no load on the shaft 20) is zero.

Waveform (e) illustrates the effect of an external steady upward force applied to the bottom of the shaft 20. This force tends to displace the piston 14 on average upwards into the upper cavity 32. The inlet port 26 is then opened for a shorter time period than the outlet port 28. The driving force F.sub.D is applied for a shorter time period as illustrated in waveform (e). The force balance condition characterized by waveforms (a) through (d) is upset and a net downward force is applied to the piston from the upper cavity 32. This net downward force due to the cavity 32 counteracts the upward external bias or load. Waveform (f) illustrates the average or net counteracting force F.sub.N which is developed during oscillator operation. These waveforms demonstrate that the hydraulic circuits act as a stiff spring to resist any average displacement of the piston and to maintain the fixed equilibrium position as illustrated in FIG. 1 about which oscillation occurs.

It is desirable that the inlet and outlet ports 26 and 28 present low impedances and that the supply and return cavities 38 and 42 also exhibit low impedances. In particular, the low impedance ports and their associated supply and return cavities 38 and 42 should present impedances which are small relative to the driving point impedance presented by the differential piston area A.sub.D, so that full supply pressure P.sub.S or full return pressure P.sub.R, is presented to the differential area 17 of the piston depending upon which port is open.

It will be noted that flow occurs both ways through the inlet port 26 and also through the outlet port 28 during each cycle of oscillation of the piston 14. While the piston is moving upward and the inlet port 26 is still open, the flow is from the supply cavity through the port 26 into the lower cavity 34. Upon the return stroke of the piston, just after the inlet port 26 is open, the flow is in the reverse direction from the lower cavity 34 through the inlet port 26 and into the supply cavity 38. A similar flow reversal occurs through the outlet port 28. Thus the low impedance cavities 38 and 42 enable the absorption of flow pulsations without causing significant changes in either P.sub.S or P.sub.R.

Under most circumstances the bulk modulus of the liquid filling the supply and return cavities 38 and 42 will enable the flow pulsations to be absorbed. In the event that the size of a liquid-filled cavity is larger than desired, a cavity arrangement as shown in FIG. 14 may be used. The illustration in FIG. 14 is for a supply cavity fed by a supply line 36 as illustrated in FIG. 1. The construction of the return cavity may be similar. The cavity is divided into a liquid-filled part 56 and a compressible gas-filled part 58; these parts being separated by a flexible diaphragm 60. The compressible gas is supplied to the part 58 through a valve 62. A perforated plate 64 supports the diaphragm 60 when hydraulic pressure is not supplied to the system (when the part 56 is not filled with pressurized fluid). The compressible gas-filled region acts as an accumulator or pressure release and enables the pressure P.sub.S in the supply cavity to remain substantially invariant with flow pulsations into and out of the cavity.

While it is desirable for the volume of the cavity 34 of FIG. 1 to be small, this cavity should have sufficient compliant reactance to reduce large pressure transients that might occur during port switching. Such pressure transients are to be avoided since they can alter the motion of the piston 14 and the valving element 16 during port switching and reduce the efficiency of hydraulic to mechanical power conversion. If additional compliance in the cavity 34 is desired, the size of the cavity can be increased as illustrated by the dash line 66 in FIG. 1.

With a much larger lower cavity, the resonant frequency of the oscillator may be made dependent both upon the stiffness presented by the liquid in the upper cavity 32 and in the enlarged lower cavity 34 when the ports 26 and 28 are both closed. In this embodiment, it may be desirable also to increase the impedance presented by the inlet and outlet ports 26 and 28 so that these ports do not short out the reactance property of the enlarged volume lower cavity 34. This may be accomplished by using less than full peripheral porting. The latter mode of porting may be accomplished by increasing the lengths of the valve sleeve 16 and providing a plurality of slots which do not occupy the full periphery of the sleeve and which cooperate with the inlet and outlet ports so as to allow a fluid to pass from these ports into or out of the lower cavity 34.

In the oscillator illustrated in FIGS. 1 through 4 the axial length of the valve sleeve may be greater than the distance between the outer metering edges 27 and 29 of the ports 26 and 28 so that the ports are not open during entire alternate half cycles of the oscillation. Such a longer valving element 16 will define a "class C" mode of operation for the oscillator.

Referring to FIG. 6, there is shown a hydroacoustic oscillator which is especially adapted for use in driving non-linear or time dependent loads. Such loads are presented by impact events. Thus the oscillator illustrated in FIG. 6 is especially adapted for use in a percussive tool in applications as rock drilling, or pile driving, or other applications where it is desired to deliver a high energy force pulse to a load. For non-symmetrical waveforms of piston motion as can occur with non-linear, or time dependent loads, port switching may best occur at other points in the cycle than the mean position as was the case for the oscillator shown in FIG. 1. It is a feature of this invention to provide, where desired, port switching at desired points in the cycle and for desired time periods. In this way an efficient conversion of hydraulic energy into high energy force pulses may readily be accomplished.

The oscillator illustrated in FIG. 6 is similar in many respects to the oscillator shown in FIG. 1, and like parts have been designated with like reference numerals. The oscillator of FIG. 6 is embodied in an impact or percussion tool where the piston 14 impacts the bottom of its shaft 20 against an anvil 68. This anvil may be part of an anvil system including an impact spring for shaping the force pulses, as for example shown in my U.S. Pat. No. 3,371,726 issued Mar. 5, 1968, or the anvil may be a drill steel, a pile cap or the top of a pile or the moil of a demolition tool.

The rings 22 and 24 are spaced from each other a distance substantially greater than the length of the valve element sleeve 70. This provides the hydroacoustic oscillator with a pressure actuated valving mechanism that achieves time delayed control over the hydroacoustic oscillator cycle. The displacement of the vlave element sleeve and the mass (piston 14) are in time delayed relationship with each other (i.e., the motion history of the valving element can be different from that of the oscillator mass.) The valve element sleeve 70 is also shown in FIGS. 7A and 7B. This sleeve 70 has a diameter larger than the diameter D.sub.S of the shaft 20. Thus the valve element sleeve 70 is freely movable on the shaft. The valving mechanism is pressure actuated in that its motion is obtained from the variations in pressure in the cavities to which the mass (piston 14) is exposed. In addition, there is a mechanism between the mass and the valving element 16 which provides for time control in the valving operations. Thus, there is provided a pressure actuated valving mechanism having means coupling the mass and valving element for controlling the timing of valving operations.

The element 70, like the element 16 (FIG. 1) has a plurality of axial passages 30 which enable the flow of fluid therethrough. Inasmuch as the valve sleeve 70 will be repeatedly engaged by and strike the upper lip of the ring 24 and the lower lip of the ring 22 as the piston 14 oscillates, it is desirable to soften these impacts. To this end notches 74 are formed at the ends of the sleeve where the sides of semi-circular passages meet. These notches 74 provide a degree of dashpot action so as to cushion the impact between the rings 22 and 24 and the ends of the valve element sleeve 70.

Unless one of the rings 22 and 24 has engaged an end of the valve element sleeve 70, the sleeve may not move. The distance separating the upper and lower lips of the rings 22 and 24 on the shaft 20, and the lengths of the valve element sleeve 70, enables the time sequence of the switching of the inlet and outlet ports 26 and 28, and the periods during which these ports are open, to be selected so as to control the frequency and force pulse delivery characteristics of the oscillator.

The operation of the oscillator illustrated in FIG. 6 may be better understood by reference to FIGS. 8, 9 and 10. FIG. 6 illustrates the condition immediately following the instant when the bottom of the shaft 20 impacts on the anvil 68. Then the valve element 70 has been driven by the ring 22 to close the outlet port 28 and open the inlet port 26. The supply pressure P.sub.S is then applied to the lower cavity 34 and acts on the differential area of the piston 14. Driving forces then applied to the piston accelerate it upwardly toward the upper cavity 32.

After a delay time T.sub.1, the piston 14 has gained substantial velocity. The lower ring 24 then engages the valve element 70, as shown in FIG. 8. Thereafter the valve element 70 closes the inlet port 26 and opens the outlet port 28. The pressure in the lower cavity 34 then drops to return pressure P.sub.R. A decelerating force is then applied to the piston 14 from the upper cavity 32. The piston continues to travel upwardly with decreasing velocity and finally reaches zero velocity at the top of its stroke after a further delay time T.sub.2. The maximum upward position of the piston is illustrated in FIG. 9.

At time T.sub.2, there is a downward force on the piston 14 due to the pressure P.sub.C acting on the area A.sub.P of the piston 14 which is exposed to the upper cavity 32. Consider, for purposes of explanation, that the return pressure P.sub.R is zero with the outlet ports 28 open as illustrated in FIG. 9. Then there is no counteracting force on the piston, and the piston is accelerated downwardly. As the piston begins to move downwardly, the lower ring 24 leaves the bottom of the valve element 70 and shortly thereafter the upper ring 22 engages the upper end of the valve element 70 as shown in FIG. 10. The valve element 70 is then driven downwardly. The piston 14 continues to pick up velocity as it travels downwardly, and impact on the anvil 68 is concurrent with maximum piston velocity. Thus, maximum kinetic energy of the piston is imparted to the anvil 68. In the latter portion of the interval of downward travel (the down stroke) of the piston 14 and concurrent with the moment of impact, the valve element 70 closes the outlet port 28 and the inlet port 26 opens. This is the initial condition for driving the piston back upwardly again, as was discussed in connection with FIG. 6.

The upper ring 22 may be adjusted so that at the instant of impact, the valve element 70 has not yet opened the inlet port 26 and closed the discharge port 28. At impact the piston will stop and the free valve element 70 may translate downward under its own inertia so as to open the inlet port 26 and close the discharge or outlet port 28 at a later time. This adjustment of the position of the ring 22 provides a time delay in the port switching action relative to time of impact so as to assure all of the kinetic energy in the piston 14 is transferred to the anvil 68, and that decelerating forces are not acting on the piston during impact.

The waveforms (a) through (d) in FIG. 11 show the variation of piston motion and forces on the piston during a plurality of successive cycles of oscillation of the piston 14 of the oscillator shown in FIGS. 6 through 10. In order to simplify the illustration as well as to show that the oscillator of FIGS. 6 through 10 can have a relaxation mode of oscillation, the waveforms of FIG. 11 show the limiting case of the dynamic stiffness of the fluid in the upper cavity 32 approaching zero so that the pressure P.sub.C in that cavity 32 remains substantially constant over the cycle and the force F.sub.SP exerted by the pressure P.sub.C on the piston area A.sub.P exposed to the cavity 32 is invariant. Waveform (b) illustrates this invariant upper cavity force F.sub.SP. In addition, the following relationships have been assumed in order to simplify the illustration.

The piston area A.sub.P is equal to twice the differential area A.sub.D, and the upper cavity pressure P.sub.C is set by the pressure divider network, including the channels of 48 and 52, the valve 50 and the restrictor 54, to be one-fourth the supply pressure P.sub.S. Accordingly, the force on the differential area F.sub.D, when port 26 is open, is equal to twice the force F.sub.SP exerted on the piston area A.sub.P.

Waveform (a) of FIG. 11 illustrates the time history of the piston displacement X.sub.P. Time T.sub.o is taken as the time immediately after impact on the anvil 68 as shown in FIG. 6. Since the inlet port 26 is open, the downward force F.sub.SP (see waveform (b)) is counteracted by the upward force F.sub.D (see waveform (c)) so that a net upward force F.sub.N (see waveform (d)) acts to drive the piston in an upward direction. The motion of the piston 14 is governed by the differential equation: ##EQU4## where a.sub.P is the acceleration of the piston and M.sub.P is the mass of the piston. At time T.sub.1 the driving force F.sub.D goes to zero, since in this illustrative example the return pressure P.sub.R is taken to be zero. The equation of motion given above reduces to ##EQU5##

The solution of these differential equations defines the piston motion X.sub.P shown in waveform (a) of FIG. 11. As graphically illustrated in that waveform, the piston moves upward in a parabolic time-displacement relationship in response to the net force F.sub.N until time T.sub.1. Then the driving force switches to -F.sub.SP. The piston continues moving upwardly in an inverse arc for a time duration T.sub.2 which is equal to T.sub.1. At T.sub.2 the velocity of the piston reaches zero. The force -F.sub.SP then drives the piston downwardly to impact upon the anvil 68 at time T.sub.P when it is travelling at maximum velocity. The solution of the differential equations given above for the illustrative case illustrated in the waveforms shows that T.sub.P is equal to 3.414 T.sub.1. The solution of these equations also shows that T.sub.1 and T.sub.2 are equal and that the piston displacements are also equal over the intervals T.sub.1 -T.sub.0 and T.sub.2 -T.sub.1.

The physical relationship of the valve element 16 and the rings 22 and 24 is that the distance separating the lower lip of the ring 22 and the upper lip of the ring 24 is equal to one-half of the peak stroke of the piston 14 plus the axial length of the valve element sleeve 70. The upper lip of the ring 24 then engages the valve sleeve element 70 on the upward stroke of the piston precisely at time T.sub.1.

It will be understood that the limiting case of invariant upper cavity force F.sub.SP and the dimensional relationships illustrated in the waveforms and in FIGS. 11 through 14 are solely for purposes of explaining an illustrative embodiment of the invention.

The upper cavity 32 may have a finite spring rate and other ratios of driving force to upper cavity force and other delays between impact and port switching T.sub.1 may be used. The separation of the rings 22 and 24 and the length of the valve sleeve element 70 may be selected to provide desired times T.sub.1, T.sub.2 and T.sub.P, as may be required by different applications for the oscillator provided by the invention.

It will be noted that waveform (d) of FIG. 11 illustrates a net downward force F.sub.N which tends to counteract the average upward thrust resulting from the average of the force pulses transferred to the anvil 68. This force F.sub.N results from the bias or restoring force developed within upper cavity 32. This average downward force F.sub.N is due to the hydraulic circuits acting as stiff springs to resist any average displacement of the piston and to maintain the equilibrium position about which the oscillation occurs.

The relaxation mode of oscillation results when the dynamic spring stiffness of the upper cavity 32 tends toward zero stiffness. This is an especially desirable mode of oscillation to precede the impact event since it makes effective use of the potential (pressure) energy stored in the upper cavity 32 over the cycle. This stored energy is effectively converted into piston motion having constant acceleration up to the instant of impact.

In the limiting case of the upper cavity 32 spring stiffness tending toward zero and for the area relationship discussed in the above given illustrative example, the frequency of oscillation is given by the expression ##EQU6## where E.sub.B is the kinetic energy of the piston on impact.

The addition of a significant dynamic spring rate to the upper cavity 32 will increase the repetition frequency of impact and change the time history of piston displacement and velocity from that shown in waveform (a) of FIG. 11.

Modifications of the upper cavity 32 so as to enable this cavity to have zero dynamic spring stiffness or other spring stiffness so as to provide said different time histories of piston displacement may be provided in accordance with this invention.

In order to provide the upper cavity 32 with a dynamic spring stiffness tending toward zero, the cavity 32 may be provided with a gas accumulator which enables the pressure to which the upper end of the piston 14 is exposed to remain substantially invariant as the piston moves over its cycle of oscillation.

As shown in FIG. 12, the upper cavity 32 may be divided into a pressurized gas-filled region 80 and a region 82, to which the piston 14 is exposed, which is filled with hydraulic fluid. This arrangement is illustrated in FIG. 12, which otherwise is similar to FIG. 6. Like parts of FIG. 12 and FIG. 6 are therefore designated by like reference numerals.

In order to provide for substantially constant supply and return pressure in the supply and return cavities 38 and 42, these cavities may also be provided with flexible diaphragms 84 and 86 which divide the cavities into two regions as was discussed in connection with FIG. 14.

Again, as discussed in connection with FIG. 14, retaining mechanisms such as the perforated member 64 shown therein may be provided for supporting the diaphragms 81, 84 and 86 when the system is not pressurized with hydraulic fluid. The gas-filled regions of the cavities 32, 38 and 40 may be filled with gas through valves 88, 91 and 93 from any suitable source of compressed gas, such as an air compressor. The provisions for gas accumulators to provide the pressure release reservoirs affords the feature of relatively small volume which can be advantageous when it is desired to provide hydroacoustic oscillators in accordance with this invention which are of minimum size.

Referring to FIG. 13, there is shown a hydroacoustic oscillator wherein the upper end of the piston 14 is exposed to the upper cavity 32 indirectly through a secondary piston 90, which has a smaller cross-sectional area than the primary piston 14. The piston 90 is free to move axially in a bore 92 which separates the bore 12, in which the primary piston 14 oscillates, from the upper cavity 32.

The region in the bore 12 between the upper end of the piston 14 and the upper end of the bore 12 is vented by means of a channel or vent 94 to the atmosphere or to drain. Accordingly, the secondary piston 90 is free to move in the bore 92 and the low pressure in the region 96 insures that the secondary piston 90 and the primary piston 14 are biased toward each other by the counteracting hydraulic pressures in the upper cavity 32 and lower cavity 34. It is desirable not to make the piston 92 integral with the piston 14 so as to facilitate manufacture of the oscillator by minimizing tolerance requirements.

The arrangement illustrated in FIG. 13 also simplifies the pressure divider arrangement for pressurizing the upper cavity 32, such that the upper cavity may be fed directly from the supply cavity 38 by means of a channel or passage 97. The passage 97 is desirably of restricted size so as not to acoustically load the system. In the exemplary case illustrated in FIG. 13, where the area of the primary piston 14 A.sub.P is twice the differential area A.sub.D and 4 times the area of the secondary piston 90, the nominal force balance in the absence of an external thrust occurs when the pressure P.sub.S in the cavity 38 is equal to the pressure P.sub.C in the upper cavity 32. The average downward force on the secondary piston 90, which has an area A.sub.C, is equal to P.sub.C A.sub.C which is equal to the average upward force on the differential area A.sub.D, or 1/2 P.sub.S A.sub.D. Thus a direct connection via the channel 97 is all that is required to establish the force balance for equilibrium. Of course the area relationships which facilitate the use of a single channel 97 need not be used and the pressure divider network illustrated in and discussed in connection with FIG. 1 or FIG. 6 may alternatively be used.

A principal feature of the oscillator illustrated in FIG. 13 is that it allows the use of an upper cavity 32 of minimum size and thus permit the oscillator system to be reduced in size on an overall basis. In other words, the arrangements illustrated in FIG. 13 results in the smallest volume of upper cavity for a given mechanical stiffness presented by that cavity to the piston 14. For the hydroacoustic oscillator illustrated in FIG. 13, the resonant frequency f.sub.o of the piston 14 is governed by the relationship ##EQU7## where V is the volume of the cavity 32, A.sub.C is the area of the secondary piston 90 which is exposed to the cavity 32 and M.sub.P is the combined mass of the primary and secondary pistons 14 and 90.

As shown in equation (9), the required volume of the cavity 32 varies directly with the square of the area A.sub.C of the piston for a given resonant frequency. Thus in the illustrated example where A.sub.P is 4 times A.sub.C, if the piston 90 were missing, as in FIG. 6, the volume V of the cavity 32 in FIG. 6 would have to be 16 times the volume of the cavity 32 in FIG. 13 for the same resonant frequency. Inasmuch as the pressure in the cavity 32 increases in the case illustrated in FIG. 13 by a factor of four over the pressures in the cavity 32 arrangements illustrated in FIGS. 1 and 6, for the same displacement of the piston 14, the stored energy in the cavity 32, which is proportional to the cavity volume and to the square of the pressure, is not altered. Thus the operating characteristics of the hydroacoustic oscillator illustrated in FIG. 13 will remain unchanged, but the size of the oscillator may be reduced.

From the foregoing description it will be apparent that there has been provided an improved class of hydroacoustic oscillators having actuation mechanisms which afford desirable modes of operation. Features of hydroacoustic oscillators which have heretofore been provided may also be used advantageously in the hereindescribed class of hydroacoustic oscillators. For example, hydrostatic bearings may be provided for lubricating the moving pistons on fluid films and rotation mechanisms for rotating anvil systems may also be provided. It will be appreciated that the foregoing description is illustrative and that variations and modifications may be provided within the scope of the invention. Accordingly, the foregoing description should be taken as illustrative and not in any limiting sense.

Claims

What is claimed is:

1. A hydroacoustic oscillator comprising

a. a movable mass having first and second sides,

b. means for continuously applying pressurized fluid to said first side to drive said mass in one direction and for storing energy due to the motion of said mass,

c. a valve mechanism having a valve element, said valve element being disposed in a path for said pressurized fluid which extends from a region of high pressure to a region of low pressure, said path being in communication with said second side, said valve element being movable in opposite directions for alternately opening said high pressure region to said second side while closing said low pressure region to said second side and then opening said low pressure region to said second side while closing said high pressure region to said second side for establishing alternating fluid pressures upon said mass to produce periodic oscillations of said mass at a frequency and over a displacement determined by said mass and the energy storage characteristics of said fluid communicating therewith, and

d. means for coupling said valve element and said mass for actuating said valve element by said means during a substantial portion of the displacement of said mass with said valve element travelling over a displacement of at least one half of the displacement of said mass so as to enable said valve to attain high velocity at the instants of the opening and closing of said second side to said high and low pressure regions.

2. The invention as set forth in claim 1 wherein said coupling means includes means for providing a portion of each cycle of oscillation of said mass during which one of said high pressure and low pressure regions is open to said second side while the other of said high pressure and low pressure regions is closed to said second side which is substantially shorter then the remaining portion of each such cycle during which said other of said high and low pressure regions is open to said second side while said one of said high and low pressure regions is closed to said second side.

3. The invention as set forth in claim 2 including a member disposed adjacent to one end of said mass for receiving impact from said end of mass during each cycle of oscillation thereof, and wherein said coupling means is spaced from said end of said mass for delaying the opening of said one of said high pressure region and low pressure region to said second side until after said impact.

4. The invention as set forth in claim 1 wherein said coupling means includes means for abruptly engaging and moving said valving element.

5. The invention as set forth in claim 1 wherein said valve element is a sleeve movably mounted on a portion of said mass which extends from said second side.

6. The invention as set forth in claim 5 wherein said coupling means includes a pair of valve element engaging means mounted on said mass portion and disposed adjacent to and spaced from the opposite ends of said sleeve.

7. The invention as set forth in claim 5 wherein said valve mechanism includes a pair of ports which are spaced from each other and are respectively in communication with said high and low pressure regions, said path extending between said ports, and wherein said sleeve extends between said ports.

8. The invention as set forth in claim 7 wherein said mass is disposed in a housing having a first cavity to which said first side of said mass is exposed, said housing also having a second cavity to which said second side of said mass is exposed, and wherein said valve mechanism is disposed in said second cavity.

9. The invention as set forth in claim 8 wherein said sleeve has passageways therethrough which extend in a direction between said ports.

10. The invention as set forth in claim 1 wherein said valving element is mounted upon and is also movable relative to said mass while said mass is in motion.

11. The invention as set forth in claim 10, wherein said coupling means includes means for transferring forces between said mass and said valving element while said mass is moving relative to said valving element.

12. The invention as set forth in claim 11 wherein said transferring means includes means for damping the motion of said valving element.

13. A hydroacoustic oscillator comprising

a. an oscillatory system including

i. a piston,

ii. a first cavity for containing pressurized fluid exposed to one end of said piston, said pressurized fluid in said first cavity continuously applying unidirectional driving forces upon said piston, and

iii. a second cavity communicating with the opposite end of said piston,

said oscillatory system having an operating frequency and displacement determined by the mass of said piston and the energy storage characteristics of the fluid communicating therewith,

b. a valving mechanism associated with said second cavity including a port structure having ports for the supply and discharge of pressurized fluid with respect to said second cavity, and

c. a valving element bi-directionally actuated by said piston for movement with said piston over at least one half of the displacement of said piston for opening and closing said ports while said valve element is traveling at high velocity to abruptly switch pressure in said second cavity between supply and discharge pressures to produce square wave driving forces on said piston.

14. The invention as set forth in claim 13 wherein said valving mechanism includes means for providing a predetermined time sequence of the opening and closing of said ports.

15. The invention as set forth in claim 14 wherein said time sequence providing means includes a lost motion mechanism coupling said piston to said element.

16. The invention as set forth in claim 15 wherein said lost motion means includes means connected to said piston for engaging said valving element being disposed on opposite sides of said element spaced from each other a distance greater than the length of said element.

17. The invention as set forth in claim 13 wherein said first cavity is spaced from said piston, and a member extending between said first cavity and the end of said piston adjacent thereto and movable with said piston in directions into and out of said first cavity, said member having a cross-sectional area exposed to said first cavity which is smaller than the cross-sectional area of said piston.

18. The invention as set forth in claim 13 wherein said first cavity is divided into a portion containing a pressurized gas and a portion containing said fluid, said fluid being a liquid, and said fluid-containing portion being exposed to said piston.

19. The invention as set forth in claim 13 including passages communicating said first cavity with supply and discharge regions through which pressurized fluid flows with respect to said second cavity, said passages defining a fluid pressure divider for providing said pressurized fluid to said first cavity and establishing an equilibrium position of said piston.

20. The invention as set forth in claim 13 wherein said first cavity is continuously filled with said pressurized fluid and continuously applies said unidirectional driving forces upon said piston.

21. The invention as set forth in claim 20 wherein said valving element is disposed bodily in said second cavity.

22. A hydroacoustic oscillator comprising

a. a housing having a bore,

b. a first member disposed for oscillatory movement in said bore,

c. said housing having a first cavity and a second cavity ddisposed on opposite sides of said first member,

d. means for maintaining said first cavity filled with fluid under pressure, for unidirectionally driving said member,

e. inlet and outlet ports, respectively, for the supply and discharge of pressurized fluid, communicating with said second cavity and spaced from each other along the path of movement of said first member,

f. a valving element disposed bodily in second cavity and bi-directionally movable by said first member to alternately open and close said inlet and outlet ports, said valving element having passage means for the circulation of fluid between opposite ends of said element, and

g. said first member having a mass and the fluid in said first cavity having energy storage characteristics which control the frequency of oscillation of said first member as a result of pressure variations in said second cavity as said valving element opens and closes said inlet and outlet ports.

23. The invention as set forth in claim 22 wherein said first member is a piston freely movable in said bore, a shaft of smaller transverse dimension extending into said second cavity from the side of said piston opposite from said side facing said first cavity and through said housing, and means on said shaft for engaging said element and moving said element.

24. The invention as set forth in claim 23 wherein said valving element is a sleeve having axial passages disposed between the inner periphery of said sleeve and the outer periphery of said shaft.

25. The invention as set forth in claim 23 wherein said inlet and outlet ports are provided by grooves in the periphery of said bore, and said valving element is a sleeve having an axial length approximately equal to the distance between the edges of said inlet and outlet port grooves which are spaced furthest apart.

26. The invention as set forth in claim 25 wherein said sleeve is freely movable on said shaft.

27. The invention as set forth in claim 26 wherein said engaging means are separated from each other a distance greater than the axial length of said sleeve.

28. The invention as set forth in claim 25 wherein said housing includes a third cavity communicating with said inlet port groove adapted to be supplied with pressurized fluid, and said housing also includes a fourth cavity communicating with said outlet port groove and a discharge channel communicating with said fourth cavity.

29. The invention as set forth in claim 28 including first and second channels each presenting a high acoustic impedance respectively communicating said first cavity with said third cavity and said first cavity with said fourth cavity to provide a pressure divider for setting the equilibrium position of said piston.

30. The invention as set forth in claim 25 wherein said first cavity includes means for reducing the stiffness presented by the fluid therein for providing a substantially constant force on said piston.

31. The invention as set forth in claim 30 wherein said last named means includes a yieldable member dividing said cavity into a liquid filled region on one side thereof facing said piston and a gas-filled region on the opposite side thereof.

32. The invention as set forth in claim 23 including a second bore in said housing between said first named bore and said first cavity, said second bore having a cross-sectional area smaller than said first bore, a second piston movable in said second bore, and means for relieving the pressure in the region of said first bore between said one end of said piston and said first cavity.

33. An impact tool for applying percussive forces to a load which comprises

a. a mass movably mounted to osciallate in directions toward and away from said load,

b. energy storage means operatively coupled to said mass for the transfer of energy therebetween,

c. means for applying pressurized fluid to said mass for applying a net force in a direction away from said load to accelerate said mass with respect to said energy storage means, and

d. means movable by said mass for controlling the application of said pressurized fluid to said mass to switch the direction of said net force alternately away from and toward said load during each cycle of oscillation of said mass with said net force being in said direction away from said load during substantially less of said cycle and toward said load for substantially more of said cycle whereby energy is transferred to said energy storage means for storage therein while said net force is in said direction away from said load to allow said stored energy to operate upon said mass when said net force is in the direction toward said load to change the direction of said mass and drive said mass to impact upon the load.

34. The invention as set forth in claim 33 when said energy storage means comprises spring means.

35. The invention as set forth in claim 34 wherein said spring means includes means reducing the dynamic spring rate of said spring toward zero while retaining a prescribed preload.

36. The invention as set forth in claim 34 wherein said spring means comprises a housing; said housing containing a first cavity in which said mass is movable and a second cavity in communication with a portion of said first cavity into which a face of said mass at one end thereof is movable, said fluid being hydraulic fluid, and means for filling said cavities with said fluid, said fluid in said second cavity providing said spring means.

37. The invention as set forth in claim 36 wherein said second cavity has a pair of chambers separated by a fluid impervious, pressure transmissive interface therebetween, one of said pair of chambers adjacent said mass face being filled with said hydraulic fluid, and the other containing a gaseous fluid.

38. The invention as set forth in claim 33 wherein said control means includes a valve element disposed on said mass and movable therewith.

39. The invention as set forth in claim 38 including means for controlling the motion of said valve element.

40. The invention as set forth in claim 39 wherein said motion controlling means includes damping means disposed between said mass and said valve element.