U.S. patent number 3,896,889 [Application Number 05/285,240] was granted by the patent office on 1975-07-29 for hydroacoustic apparatus.
This patent grant is currently assigned to Hydroacoustics, Inc.. Invention is credited to John V. Bouyoucos.
United States Patent |
3,896,889 |
Bouyoucos |
July 29, 1975 |
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.
Inventors: |
Bouyoucos; John V. (Brighton,
NY) |
Assignee: |
Hydroacoustics, Inc.
(Rochester, NY)
|
Family
ID: |
23093390 |
Appl.
No.: |
05/285,240 |
Filed: |
August 31, 1971 |
Current U.S.
Class: |
173/120; 91/52;
91/276; 91/321; 92/134; 175/56 |
Current CPC
Class: |
F01B
11/06 (20130101); F01L 21/04 (20130101) |
Current International
Class: |
F01B
11/06 (20060101); F01B 11/00 (20060101); F01L
21/04 (20060101); F01L 21/00 (20060101); F01l
017/00 (); F01b 007/18 () |
Field of
Search: |
;91/276,218,52,321,443
;175/56 ;173/116,119,120 ;92/134 |
References Cited
[Referenced By]
U.S. Patent Documents
Foreign Patent Documents
Primary Examiner: Maslousky; Paul E.
Attorney, Agent or Firm: Lukacher; Martin
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.
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.
* * * * *