U.S. patent number 3,768,576 [Application Number 05/187,299] was granted by the patent office on 1973-10-30 for percussion drilling system.
Invention is credited to Leo A. Martini.
United States Patent |
3,768,576 |
Martini |
October 30, 1973 |
PERCUSSION DRILLING SYSTEM
Abstract
A unique, variably damped, steady state, broad range, fluid
forced, force superposition, oscillatory, percussion drilling
method and means, capable of resonant frequency operation and force
magnification, is disclosed that is operable from liquid or gaseous
fluid under pressure for rotary drilling of oil, gas, and water
wells, geophysical holes, open strip mining blast holes,
construction holes, and the like for greatly increasing the rate at
which said bore holes are drilled. This comparatively simple prime
mover produces sustained high frequency, high amplitude,
longitudinal force spikes on a drill bit and synchronizes
consumption of percussive force energy and drill collar weight
force energy and superimposes one force upon the other to obtain
instantaneous anvil accelerations of much greater magnitude than
either force could, acting separately, to produce rock crushing
forces of greater effectivity. In addition to force superposition
an inphase energy consumption, this device when operating at or
about resonant frequency will produce a phenomenon of force
magnification, force reinforcement, and maximum
transmissibility.
Inventors: |
Martini; Leo A. (Dallas,
TX) |
Family
ID: |
22688408 |
Appl.
No.: |
05/187,299 |
Filed: |
October 7, 1971 |
Current U.S.
Class: |
173/73; 175/56;
175/92; 173/136 |
Current CPC
Class: |
E21B
4/14 (20130101); E21B 7/24 (20130101) |
Current International
Class: |
E21B
7/24 (20060101); E21B 7/00 (20060101); E21B
4/14 (20060101); E21B 4/00 (20060101); E21b
001/06 (); E21b 005/00 () |
Field of
Search: |
;175/56,92
;173/73,80,133 |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Leppink; James A.
Parent Case Text
This application is a continuation-in-part of application Ser. No.
18,635, filed Mar. 11, 1970, now U.S. Pat. No. 3,612,191, granted
Oct. 12, 1971.
Claims
What is claimed is:
1. A percussion drilling apparatus comprising:
a drilling mass having a fundamental damped vibrational
frequency;
means for elastically biasing the drilling mass into engagement
with a formation to be drilled;
driver means for reciprocating the drilling mass toward the
formation at a rate substantially equal to its fundamental damped
vibrational frequency; and
fluid means for actuating the driver means to reciprocate the
drilling mass.
2. A prime mover apparatus for drilling purposes comprising:
a drilling mass;
elastic restoring means for cooperation with the drilling mass to
provide a natural damped period of oscillation of the drilling
mass, and for biasing the drilling mass toward a formation to be
drilled;
fluid powered hammer means for reciprocating the drilling mass
toward the formation at a frequency substantially identical to the
natural damped period of oscillation of the drilling mass; and
valving means for regulating the operation of the fluid powered
hammer means.
3. A forced resonant harmonic prime mover apparatus for borehole
drilling comprising:
a reciprocal drilling mass;
means for elastically restoring the mass to provide a natural
damped period of oscillation of the mass and for biasing the
drilling mass toward a formation to be drilled;
means for reciprocating the drilling mass relative to the formation
at a frequency substantially equal to the natural period of
oscillation of the mass; and
fluid means for actuating the means for reciprocating the drilling
mass at a frequency substantially equal to the natural period of
oscillation of the drilling mass and thereby reinforcing the
movement of the mass to increase the amplitude of oscillation of
the mass.
4. A forced resonant harmonic prime mover for borehole drilling
according to claim 3 wherein the means for elastically restoring
the drilling mass includes means for applying a separate force
directed along the axis of oscillation of the drilling mass and
thereby superimposing the oscillation forces of the drilling mass
on the separate force.
5. The prime mover according to claim 4 wherein the means for
reciprocating the drilling mass includes a reciprocal hammer
powered in both directions of reciprocation by fluid under
pressure.
6. A percussion drilling apparatus comprising:
a down hole reciprocatory percussion motor;
a drilling bit mounted for outward reciprocation by the percussion
motor to form a borehole in the earth;
a drill string for supporting the percussion motor and the drilling
bit and for applying an outwardly directed force; and
spring means mounted between the drill string and the drilling bit
for applying the drill string force to the drilling bit during the
outward reciprocation of the bit by the percussion motor so that
the outwardly directed force of the percussion motor is
superimposed on the outwardly directed force of the drill string to
increase the drilling action of the bit.
7. The percussion drilling system according to claim 6 wherein the
percussion motor reciprocates the drilling bit at a predetermined
frequency and wherein the means for applying the force of the drill
string to the drilling bit comprises at least one elastic element
mounted between the drill string and the drilling bit for
cooperation with the formation of the earth to oscillate the
drilling bit at a frequency substantially identical to the
predetermined frequency of operation of the percussion motor.
8. The percussion drilling system according to claim 7 wherein the
percussion motor comprises an anvil slidably supported at the
distal end of the drill string and having the drilling bit secured
thereto, a hammer mounted for reciprocation into and out of
engagement with the anvil, and valving means for alternately
directing fluid pressure supplied through the drill string to the
opposite sides of the hammer and thereby causing alternate inward
and outward reciprocation of the hammer.
9. The percussion drilling system according to claim 8 wherein the
valving means comprises a valving member mounted on the hammer for
movement with respect thereto under its own inertia to a first
valving position whenever the hammer reaches one extreme of its
reciprocation and for movement under its own inertia to a second
valving position whenever the hammer reaches the other extreme of
its reciprocation, and means for moving the hammer toward the anvil
whenever the valving member is in the first valving position and
for moving the hammer away from the anvil whenever the valving
member is in the second valving position.
10. The percussion drilling system according to claim 8 wherein the
valving means comprises a member extending from the drill string
through the hammer to the anvil and having pressurized fluid inlet
and exhaust fluid outlet passages formed in it, and a pair of
passageways formed through the hammer, the first extending to one
end of the hammer and including a portion located for communication
with the inlet passageway when the hammer is at one extreme of its
reciprocation and for reciprocation with the outlet passageway when
the hammer is at the other extreme of its reciprocation, and the
second extending to the opposite end of the hammer and including a
portion located for communication with the inlet passageway when
the hammer is at one extreme of its reciprocation and a portion
located for communication with the outlet passageway when the
hammer is at the other extreme of its reciprocation.
11. A percussion drilling system comprising:
a drilling bit for penetrating a formation of the earth in a
predetermined direction;
drill string means providing a force having the predetermined
direction;
means for reciprocating the drilling bit relative to the drill
string means; and
means mounted between the drill string means and the drilling bit
for transferring the drill string force energy to the drilling bit
inphase with the movement of the bit toward the formation under the
action of the reciprocating means.
12. The percussion drilling system according to claim 11 wherein
the force energy transferring means further comprises means for
storing energy from the force of the drill string during movement
of the bit toward the drill string.
13. The percussion drilling system according to claim 12 wherein
the force energy transferring means comprises at least one spring
mounted between the drill string means and the drilling bit and
having such a spring constant that the formation of the earth and
the spring cooperate to restore the drilling bit substantially
inphase with the movement of the drilling bit reciprocating
means.
14. The percussion drilling system according to claim 13 wherein
the drilling bit reciprocating means comprises a percussion motor
mounted in the distal end of the drill string and including a
hammer which is reciprocated both inwardly and outwardly relative
to the drill string means under the action of fluid pressure.
15. A system for drilling a borehole in a formation of the earth
comprising:
a drill string;
a drilling anvil-bit mass mounted at the distal end of the drill
string;
means for cooperation with the formation being drilled and the
anvil-bit mass to oscillate the anvil-bit mass at a natural damped
period of oscillation; and
means for reciprocating the drilling anvil-bit mass outward during
outward oscillation of the anvil-bit mass to reinforce the outward
oscillation of the anvil-bit mass and thereby magnify the anvil-bit
oscillation.
16. The system according to claim 15 wherein the means for
cooperation with the formation being drilled and the anvil-bit mass
includes means for applying the force of the drill string and
thereby superimposing the anvil-bit mass oscillation forces on the
force of the drill string.
17. The system according to claim 16 wherein the drill string
provides an outwardly directed force and wherein the means for
cooperation with the formation being drilled and the anvil-bit mass
includes means for storing the force energy of the drill string
during inward oscillation of the anvil-bit mass and transferring
the force energy of the drill string to the anvil-bit mass
simultaneously with the outward oscillation of the anvil-bit mass
and thereby changing the phase of energy application from the drill
string to the anvil-bit.
18. The percussion drilling system according to claim 15 wherein
the means for reciprocating the anvil-bit mass includes a
reciprocating hammer and wherein the anvil-bit mass oscillating
means comprises at least one spring mounted between the drill
string and the anvil-bit mass which in combination with the
formation being drilled forms a combined elastic system that
functions to substantially return the anvil-bit mass to a
predetermined location in a prescribed period of time with a
predetermined direction of movement.
19. The drilling system according to claim 18 further comprising an
anvil as part of the bit mass for impact by the hammer and wherein
the hammer is driven in both directions of its reciprocation by
fluid pressure which is supplied through the drill string.
20. A percussion drilling system according to claim 19 wherein the
drilling anvil-bit mass is retained in one directional longitudinal
movement with the drill string by a retainer in the groove of the
drilling mass having an outer surface larger than the reduced
internal surface of the drill string and having a surface that
coacts with the reduced internal surface of the drill string so
that when said surfaces are forced together said surfaces will
forcibly keep the retainer in the groove of the drilling anvil-bit
mass.
21. The borehole drilling system according to claim 15 wherein the
anvil-bit reciprocating means comprises a hammer mounted for
reciprocation into and out of engagement with the anvil-bit and
valving means for alternately directing fluid pressure supplied
through the drill string to the opposite sides of the hammer and
thereby causing alternate inward and outward reciprocation of the
hammer.
22. The borehole drilling system according to claim 21 wherein the
valving means comprises a valving member mounted on the hammer for
movement with respect thereto under its own inertia to a first
valving position whenever the hammer reaches one extreme of its
reciprocation and for movement under its own inertia to a second
valving position whenever the hammer reaches the other extreme of
its reciprocation, and means for moving the hammer toward the
anvil-bit whenever the valving member is in the first valving
position and for moving the hammer away from the anvil-bit whenever
the valving member is in the second valving position.
23. The borehole drilling system according to claim 21 wherein the
valving means comprises a member extending from the drill string
through the hammer to the anvil-bit and having pressurized fluid
inlet and exhaust fluid outlet passages formed in it, and a pair of
passageways formed through the hammer, the first extending to one
end of the hammer and including a portion located for communication
with the inlet passageway when the hammer is at one extreme of its
reciprocation and for reciprocation with the outlet passageway when
the hammer is at the other extreme of its reciprocation, and the
second extending to the opposite end of the hammer and including a
portion located for communication with the inlet passageway when
the hammer is at one extreme of its reciprocation and a portion
located for communication with the outlet passageway when the
hammer is at the other extreme of its reciprocation.
24. A drilling prime mover consisting of:
an anvil-bit mass;
pipe means supporting the anvil-bit mass;
means mounted between the pipe means and the anvil-bit mass for
cooperation with the formation being drilled and the anvil-bit mass
to cause oscillation of the anvil-bit mass at its fundamental
damped vibrational frequency; and
means for periodically impulsively forcing the anvil-bit in its
direction of movement during oscillation to increase the amplitude
of anvil-bit oscillation.
25. The prime mover according to claim 24 wherein the pipe means
provides a force and the means mounted between the pipe means and
the anvil-bit mass for cooperation with the formation being drilled
and the anvil-bit mass includes means for transferring the force of
the pipe means so that the oscillation forces of the anvil-bit are
superimposed on the force of the pipe means.
26. The prime mover according to claim 24 wherein means mounted
between the pipe means and the anvil-bit mass is an elastic member
that biases the pipe means and the anvil-bit mass in opposite
directions.
27. A method of forming a borehole by percussion drilling including
the steps of:
storing energy in at least one spring positioned between a drill
string and a drilling bit mounted at the distal end of the drill
string;
periodically forcing the bit outwardly from the drill string and
into engagement with a formation to be drilled; and
transferring the energy stored in the spring to the bit inphase
with the outward movement of the bit and thereby increasing the
drilling action of the bit.
28. The method according to claim 27 wherein the step of forcing
the bit outwardly from the drill string is carried out by
repeatedly actuating the bit with a percussion motor.
29. The method according to claim 28 wherein the step of forcing
the drilling bit outwardly from the drill string is further
characterized by repeatedly impacting an anvil connected to the
drilling bit with a reciprocatory hammer which is actuated by fluid
pressure supplied through the drill string.
30. The method according to claim 27 wherein the bit has a natural
period of oscillation and further characterized by forcing the bit
outwardly at a frequency substantially equal to the natural period
of oscillation of the bit.
31. A method of percussion drilling comprising:
oscillating a drilling bit at a predetermined natural period of
oscillation; and
impacting the drilling bit at a frequency which is substantially
matched to the predetermined natural period of oscillation.
32. The method of percussion drilling according to claim 31 wherein
the oscillating step is carried out by biasing the drilling bit
with a spring which cooperates with a formation to be drilled to
oscillate the drilling bit at said natural period of
oscillation.
33. The method of percussion drilling according to claim 32
including the additional step of storing energy in said spring and
transferring energy from said spring to the drilling bit during
outward reciprocation of the drilling bit.
34. The method of percussion drilling according to claim 33 wherein
the impacting step is carried out by impacting the bit with a
reciprocating hammer which is driven both into and out of
engagement with the drilling bit under the action of fluid
pressure.
35. A method of drilling a borehole including the steps of:
storing energy in an elastically biased mass vibrational system
having a natural period of vibration; and
periodically adding energy to the oscillating mass at a frequency
substantially equal to the natural period of vibration of the mass
to increase the energy of the mass.
36. The method of drilling a borehole according to claim 35
including the additional step of adding additional energy to the
oscillating mass by forcing the mass along the axis of
oscillation.
37. A method of drilling a borehole according to claim 36 wherein
the step of periodically adding energy to the oscillating mass is
carried out by the percussive impact of a reciprocal hammer driven
in both directions of reciprocation by pressurized fluid.
38. A percussion drilling apparatus comprising:
a drill string forming a pressurized fluid inlet passageway;
an anvil-bit slidably supported at the distal end of the drill
string and including a fluid discharge passageway;
means forming a cylinder adjacent the anvil-bit;
a piston mounted in the cylinder for reciprocation therein into and
out of engagement with the anvil-bit;
said piston having at least two passageways formed through it, the
first extending into communication with the portion of the cylinder
beneath the piston and the second extending into communication with
the portion of the cylinder above the piston;
fluid inlet means for directing pressurized fluid from the drill
string to a pressurized fluid outlet opening; and
fluid outlet means for directing exhaust fluid from an exhaust
fluid inlet opening to the fluid discharge passageway of the
anvil-bit,
valving means responsive to positioning of the piston in the lower
portion of the cylinder to connect the pressurized fluid outlet
opening in communication with the first passageway in the piston to
direct pressurized fluid to the portion of the cylinder beneath the
piston and to connect the exhaust fluid inlet opening in
communication with the second passageway in the piston to direct
exhaust fluid from the portion of the cylinder above the piston
into the fluid discharge passageway of the anvil-bit and responsive
to positioning of the piston in the upper portion of the cylinder
to connect the pressurized fluid outlet opening in communication
with the second passageway in the piston to direct pressurized
fluid to the portion of the piston above the cylinder and to
connect the exhaust fluid inlet opening in communication with the
first passageway in the piston to direct exhaust fluid from the
portion of the cylinder beneath the piston into the fluid discharge
passageway of the anvil-bit.
39. A percussion drilling system according to claim 38 wherein the
fluid inlet means and the fluid outlet means comprise passageways
formed in a member mounted in the cylinder and extending from the
drill string through the hammer to the anvil-bit.
40. The percussion drilling system according to claim 39 further
including a reduced diameter passageway extending through the
member from the passageway in the member comprising the fluid inlet
means to the passageway in the member comprising the fluid outlet
means for providing continuous fluid flow through the fluid
discharge passageway of the anvil-bit while maintaining a pressure
differential between the fluid inlet passageway of the drill string
and the fluid discharge passageway of the anvil-bit.
41. The percussion drilling system according to claim 40 wherein at
least portions of the passageways in the piston are formed in a
valve block mounted in the piston for reciprocation therewith.
42. In a percussion drilling system of the type including an
anvil-bit having a fluid discharge passageway formed through it, a
hammer mounted for reciprocation into and out of engagement with
the anvil-bit, and a drill string for supporting the anvil-bit and
the hammer and having a pressurized fluid inlet passageway formed
through it, the improvement comprising:
a member extending from the drill string through the hammer to the
anvil-bit, said member having a pressurized fluid passageway
extending through it from the pressurized fluid inlet passageway in
the drill string to a pressurized fluid outlet opening and having
an exhaust fluid passageway extending through it from an exhaust
fluid inlet opening to the fluid discharge passageway in the
anvil-bit; and
two passageways formed through the hammer, the first extending to
one end of the hammer and including a portion located for
communication with the pressurized fluid outlet opening when the
hammer is at one extreme of its reciprocation and another portion
located for communication with the exhaust fluid inlet opening when
the hammer is at the other extreme of its reciprocation, and the
second extending to the opposite end of the hammer and including a
portion located for communication with the pressurized fluid outlet
opening when the hammer is at one extreme of its reciprocation and
another portion located for communication with the exhaust fluid
inlet opening when the hammer is at the other extreme of its
reciprocation.
43. The improvement according to claim 42 wherein the pressurized
fluid passageway and the exhaust fluid passageway of the member are
of relatively large diameter and further including a relatively
small diameter passageway extending through the member from the
pressurized fluid passageway to the exhaust fluid passageway for
permitting constant fluid flow through the exhaust fluid passageway
to the fluid discharge passageway of the anvil-bit while
maintaining a predetermined pressure differential between the
pressurized fluid passageway and the exhaust fluid passageway.
44. The improvement according to claim 43 wherein the pressurized
fluid outlet opening and the exhaust fluid inlet opening are
outwardly facing and wherein the portions of the passageways of the
hammer located for communication with the pressurized fluid inlet
opening and the exhaust fluid outlet opening are inwardly facing so
that sliding valve action is effected during reciprocation of the
hammer.
45. For use in a percussion drilling apparatus, an anvil-bit having
a mass which cooperates with elastic restoring forces to establish
a natural period of oscillation of the anvil-bit and having a
drilling bit on one end thereof, an anvil on the opposite end
thereof, a shank diameter smaller than the bit extending from the
anvil end, rotational drive surfaces on said shank, a retainer
groove near said anvil end on said shank, fluid passageway through
said anvil, shank, and bit and a substantially radially surface
connecting the outside of the shank surface with an enlarged
diameter of the anvil-bit mass extending from the drilling bit end
toward the anvil end and adapted to receive the application of said
elastic restoring forces.
Description
The invention encompasses first a forced, damped-spring-mass
oscillatory vibrational system formed in part by the percussion
tool and formed in part by the formation being drilled and these
parts cooperate to determine in part the operational
characteristics of the drilling system. Secondly, the invention
involves a system exciting fluid driven percussive motor which may
use either of two automatic valving structures, one for liquid and
the other for gaseous pressurized circulating mediums thereby
providing for multiple industry application.
BACKGROUND OF THE INVENTION
Although the concept of percussion drilling is old, the need exists
today more than ever before for a tool that satisfies the rigorous
requirements of a truly effective and successful tool. With wells
being drilled deeper and into hard virtually inpenetrable earth
strata, it is imperative that improvements over existing devices be
made since generally today they are still lacking in
effectivity.
Despite many frustrating failures of man and machine in this area,
more intense liquid percussion tool development has spanned this
decade, and has progressed substantially as new insights occurred,
and new materials, processes, and techniques were proven and
adopted. These devices are progressing toward market and promise to
be in general usage in a few years.
In the past, emphasis has been on basic approaches to rock attack,
compatability with existing equipment and systems, suitable tool
operation, valve erosion prevention and other fundamental design
parameters. With many of these things having been determined, the
effort has shifted to increasing tool effectivity, increasing
service life of major as well as minor components, making tools
field serviceable and reducing manufacturing costs to make them
economically justifiable.
The trend today is to power generation for drilling at or near the
bit where the work is to be done because of several factors. One is
due to power losses and considerable drill pipe wear incurred
mostly from the pipe string rubbing the borehole wall between the
bit and the generally remote surface locations, ranging today up to
and above 20,000 feet deep as in oil and gas wells. Another factor
in this trend is that bit life per foot of hole drilled can be
increased when drilling with a percussion tool, thereby reducing
the number of times the drill pipe has to be pulled out of the
borehole to replace the bit. Another contributing factor is that
straighter boreholes are possible, thus diminishing the chances of
troubles with stuck or twisted off drill pipe, well casing
installation, and well pumping unit problems. The major factor,
however, influencing the endeavor to produce more effective
percussion tools by today's standards, is the promise of well
completion in less time with the associated reduction of labor,
material, and equipment costs.
Since over 90 percent of the oil and gas wells being drilled today
use liquid, usually water or an oil-based mud, as the
system-circulating medium, tool operation on said liquid is
naturally the area of greatest concentrated effort but also harder
to achieve because of the inelasticity or non-compressible
characteristics of the liquid and the circulated solids that pass
through the tool causing erosion and wear. Generally the operating
conditions of this type tool are severe, considered from every
aspect, and historically have not lived up to the required
standards of being reasonably long lasting, having the desired
effectivity, and being comparatively simple in construction.
AMONG PRIOR ART CONSIDERED
U. S. Pat. No. 1,892,517--Pennington--Dec. 27, 1952
U. S. Pat. No. 2,774,334--Cline, Jr.--Dec. 18, 1956 18,1956
U. S. Pat. No. 2,859,733--Bassinger et al.--Nov. 11, 1958
U. S. Pat. No. 3,410,353--Martini--Nov. 12, 1968
U. S. Pat. No. 3,327,790--Vincent et al.--June 18, 1967
U. S. Pat. No. 3,387,671--Collier--June 11, 1968
SUMMARY OF THE INVENTION
This invention relates to a dual operational mode, fluid driven,
forced frequency, impulsive force, steady state, variably damped
oscillator capable of percussive force and superposition force
production in the stiffness controlled operational mode and capable
of producing percussive force, superposition force, and force
magnification in the mass controlled resonant frequency mode.
This invention relates to prime movers of the reciprocating impact
type and, more particularly, to a fluid actuated percussion
drilling tool capable of converting fluid energy to mechanical
energy and to a means and method of increasing the effectivity and
power generation of same and having improved construction and
operating characteristics.
A primary object of this invention is to provide a prime mover of
the reciprocating type that will maintain drill collar weight force
on a drill bit during and after the time percussive impact force is
applied, thereby adding or superimposing said forces, one on the
other for more effective formation cleavage.
Another primary object of this invention is to provide a prime
mover that will maintain drill collar weight force on a drill bit,
superposition percussive blow force on said weight force and
reinforce said forces with other generated inertial mass forces
produced at resonant frequency system operation.
An object of this invention is to provide a stiffness controlled,
forced, oscillator with critical or overdamping characteristics
whereby superposition percussive blow forces are exerted in phase
with other applied forces during nonresonant frequency
operation.
Another important object of this invention is to provide a mass
controlled resonant frequency oscillator that during operation will
store or expend energy relative to the damping effect of the system
output and will produce force magnification relative to the output
work.
Still another important object of this invention is to provide an
oscillatory percussion tool that will operate effectively in either
mode of resonant and non-resonant frequencies and be capable of
operational mode change.
Another object is to provide a device of this type that will
generate more mechanical power output for a given power input by
conserving system energy and a device that will automatically
adjust its operation according to the formation being drilled.
Another object is to provide a reciprocating impact prime mover
particularly well suited to operation from a relatively
incompressible pressure fluid that incorporates among other
desirable features an improved valving means that lies within the
hammer, is rapidly shiftable therewith precisely at termination of
hammer stroke, and has one shiftable valve member that
simultaneously opens one set of valve surfaces while closing
another set of valve surfaces, and cooperates with the hammer to
provide a true "on demand" cycling mechanical hydraulic system.
Another primary object of this invention is to provide a percussion
drill that is comparatively small in size, is comparatively
economical to manufacture, has increased drilling effectivity, and
has sufficient durability that it can be used in most all drilling
operations, thereby reducing well completion time with the
associated reduction of labor, material and equipment costs.
An object of this invention is to make more effective use of drill
collar weight when drilling with a percussion motor.
An object of this invention is to provide a drilling system that
will cooperate with the formation being drilled to determine in
part the operating characteristics of the system.
Yet another object of this invention is to provide two separate
fluid valving constructions for the percussion motor which excites
the vibrational system that will allow more optimized performance
and economy for each of the fluid states, gaseous and liquid.
Yet another object of this invention is to provide a special
application device that is capable of drilling out well control
tools such as cementing apparatus, bridge plugs, and packers that
may be made of aluminum, magnesium, cast-iron, or other
materials.
The foregoing objects, together with other objects, will be more
fully apparent from the descriptions that follow.
In the drilling industry standard arrangement where a percussion
tool is installed in the drill string with a bit attached at its
lower end and drill collars attached above the tool, this is
generally assumed to be the best application of drilling technique.
It, however, has proven to be less effective than anticipated with
the percussion tools known today. With use of better test
instruments and greater insight into what truly happens during
operation of the drilling system, it has been found that while all
known percussion tools produce a vibratory movement or oscillate
axially and while some worthwhile increase in penetration rate is
produced, it is far from being the most advantageous. On
conventional tools of this type it was found that at percussion of
the hammer with the anvil-bit, the anvil-bit is impulsively driven
forward relative to the weight force transferring surface of the
drill string and out of engagement with same, and that during said
disengagement the drill string weight force is released from the
anvil-bit. In other words, drill collar weight will come to rest on
the anvil before the percussive blows but at percussion, the anvil,
due to the percussive force, is driven out from under the drill
collar weight. This is due to the fact that the percussive blow is
very fast, on the order of 0.0002 second, and in this increment of
time the anvil can achieve an instantaneous acceleration rate many
times the gravitational acceleration rate of a free falling body,
such as that of the drill collar weight, regardless of its mass
and, thus, cannot maintain its weight force on said anvil. Further
examination of tools in use today determined that the drill collar
weight force energy application was approximately 180.degree. out
of phase with the percussive blow. That is, the percussive blow
energy is applied and used from 0.degree. to 180.degree. with the
maximum anvil displacement occurring at 180.degree.. The following
but slower drill collar load is then applied and comes to rest on
the anvil-bit applying and using its energy after 180.degree. and
before 360.degree. which indicates that it is completely out of
phase to produce additive forces or be of much value in drilling.
In fact this condition is considered highly undesirable. Drill
collar weight then does not enhance percussion tool operation as it
has been applied, but merely produces an interchange of percussive
force and drill collar weight alternately applied to the bit with
little advantage. Percussive blow force actually offsets or negates
the drill collar weight that would normally be applied to the bit,
because force applications are out of phase and tend to cancel each
other. It should be noted that the drill collar weight is not
maintained on the anvil and the bit, and, particularly, that no
anvil and bit accelerations are possible due to drill collar weight
during percussion in the situation described. When percussive blow
force is maximum on the bit, drill collar weight force on the bit
is essentially zero and when drill collar weight force on the bit
is maximum, percussive blow force on the bit is zero and so
uncoordinated, unsynchronized forces come into play producing a
vibrating action of lesser effectivity.
To synchronize and substantially superposition percussive blow
force and drill collar weight energy applications, extensive
mechanical force storage and energy phase transfer means in the
form of anvil thrust rings have been provided on the percussion
tool between the driver sub and the anvil shoulder. These anvil
thrust ring springs store energy as they are compressed and release
energy as they expand and thereby change the phase of drill collar
energy usage from approximately 270.degree. to 0.degree. to
180.degree. which is inphase with the percussive blow energy usage.
This delayed force energy can follow and add to instantaneous
percussive force and anvil accelerations after percussion. This
arrangement prevents cancellation of percussive force by reduction
of drill collar weight force and will achieve increased bit
effectivity.
The anvil thrust rings may be thick, high force, low inertia, short
reaction time frustro-conical or other type ring or coil springs as
manufactured and will exert forces in resistance to loadings
tending to flatten, deflect, or distort them and will exert said
forces through their shape change. They compress, flatten or
distort under drill collar load and transfer said load to the bit.
In the situation as before percussion, force on the bit is drill
collar load force, but at percussion, force on the bit is drill
collar weight force as applied by anvil thrust rings plus
percussive blow force. It is the objective here to stack these
forces one on the other at percussion to make them more useful
rather than let them be cancelled at percussion. Specifically in
the case of a critically damped system, potential energy is stored
in the anvil thrust rings after the maximum anvil-bit displacement
which is 180.degree. through the rest of the tool cycle back to
0.degree. at which time another percussive blow occurs. At the
percussion blow when threshold rock crushing forces are achieved by
a combination of drill collar and percussive load these energies
are used and consumed together from 0.degree. to 180.degree..
Energy usage is then in phase and forces are additive. Drill collar
weight force energy applied between percussions is stored for use
at percussions, thus bringing these forces inphase to increase the
force amplitude. Anvil acceleration forces at percussion can be
greatly increased since drill collar weight energy stored in the
anvil thrust rings is also applied at percussion and will produce
high magnitude rock crushing forces. The instantaneous high forces
become very worthwhile and effective when drilling the harder
formations and crushing forces over a distance need to be applied
for good bit tooth penetrations.
The anvil thrust rings also provide other desirable features such
as allowing the drill collars to move downward more smoothly and
uniformly, tending to cushion their movement as energy is absorbed
by the anvil thrust rings and thereby isolating the drill string
above from rapid anvil movements. They also tend to stabilize the
torque required to rotate the drill stem and hence reduce dynamic
torsional stress in the drill string.
The simple foregoing explanations are pertinent and relevant and
form an important part of this specification but the invention
encompasses more than the incorporation of the anvil thrust ring
springs, and new valving structures. It is a new method and means
for down the hole drilling system operation.
Due to the discovery that drill collar weight as being used was for
the most part ineffective for increasing the rate of penetration
when used with a percussive motor, an investigation of the drilling
system as well as the formation being drilled was undertaken in an
effort to improve system effectivity. Out of these studies came
significant data which will surely advance drilling technology.
Although much is yet to be learned about the visco-elastic
properties of rock, formation elasticity under various hydrostatic
pressures, rock fracture time versus elasticity, and overburden
stress effects, certain data is known and form in part the basis of
this invention. The research indicates that most all formations
show greater compressive strengths with increased confining
pressure. Under increased confining pressures formations exhibit
large amounts of deformation and enhanced ductility prior to
failure. Rock with brittle viscous behavior at atmospheric
pressures will show remarkable elasticity under the forces of
hydrostatic well pressures and heavy earth overburden and these
physical characteristics of rock under stress will have a
prounounced influence on the energy requirements for fracture by
action of bit teeth and the resulting volume of dislodged
formation. It was noted that the springlike condition of the
formation was capable of storing a portion of the energy from a
percussive blow and returning said energy to the anvil-bit and that
this energy could be utilized in conjunction with the existing
anvil thrust rings to increase the oscillation amplitude of the
drilling bit and ultimately enhance system performance. This
apparent natural obstacle of formation elasticity to drilling
became an asset, was incorporated, and used to good advantage for
conservation of system energy and by providing a unique tool that
is responsive to formation conditions. It resulted in a drilling
tool that will perform as well in deep bore holes as on the surface
and in plastic, rubbery, tough, hard, elastic brittle formations,
and one that can automatically adjust its force level
accordingly.
To more fully explain the salient features embodied in this
invention a technical and theoretical explanation of system
operation is required. In this way a clear presentation of concepts
and principles involved can be made and differences from existing
systems can be viewed with respect to novelty and invention.
In a broader view this invention encompasses, first, the formation
to be drilled, a mass that can be considered infinite, of varying
hardness and elasticity, having certain energy absorbtion and
spring rate characteristics and having certain damping qualities.
Secondly, the invention involves a method and means of expediently
penetrating the formation, having a part of comparatively large
mass relative to the other parts and consists of the drill string.
On the distal end of this drill string is an anvil-bit with
acertain weight or mass and suitably constructed to mate with the
drill string on its upper side and to engage and dislodge the
formation on its lower side. Interposed between the anvil-bit and
the drill string are the anvil thrust ring springs, having a
suitable spring rate and certain small inherent damping qualities.
The anvil-bit is mounted for longitudinal oscillation relative to
the drill string and the formation as a whole. A reciprocal hammer
mass motor located inside the drill string casing is provided to
cause the anvil-bit oscillation in its one degree of freedom and is
responsive to the pressurized fluid stream in said drill string.
The system is shown schematically in FIG. 17. This drilling system
then embodies a vibrational mass under the influence of restoring
forces that is adapted to be excited by a series of periodic
impulses defining a forced oscillator and when certain conditions
exist is a forced resonant frequency oscillator. These certain
conditions are a particular relationship between the combined
system restoring forces and the oscillatory mass and also having
another particular relation with the periodic impulses. The forced
stiffness controlled oscillator and the forced mass controlled
resonant oscillator produce system output forces that are
superpositioned on any preload such as drill collar weight that the
system may have and the energy the oscillator produces is consumed
simultaneously with the preload energy.
The formation effect on the drilling system described herebefore is
descriptive of the inherent natural internal behavior of rock under
various stress effects and although the elastic properties and
potential energy storage therein contribute to the elastic
restoring forces of the vibrational system, they make lesser
overall contributions to the system than the greater governing
physical laws of mechanics under which the system functions. These
physical laws produce a system condition that can be called
spring-like or as having the nature of an elastic member wherein a
force is present, a displacement is present and a spring rate is
produced in that there is a force change per unit displacement as
well as potential energy storage relative to the resisting force.
This is one aspect that in cooperation with the other parts allows
this oscillatory and harmonic vibrational system to perform with
elastic restoring forces on the anvil-bit and thus a true spring
mass system even through the formation disintegration beneath the
bit may be totally of a viscous damping nature.
The earth as a whole and the drill string that comes to bear on
said earth have a mutual gravitational attraction relative to the
mass of each. This said attraction causes equal and opposite forces
at the lower end of the drill string on the bit face and the
oppositely facing formation face in contact with the bit. The
formation then is in effect pushing back against the bit as much as
the bit is pushing against the formation. As the anvil-bit is
driven forward by the percussive blow, damping occurs as the
formation disintegrates before it but as the energy is consumed,
increasing resistance builds up to the bit which is then
momentarily halted to forward movement. But since the average rate
of forward travel of the bit is the same as the travel of the drill
string, when the anvil-bit stops its forward movement, it
effectively makes a reverse stroke in that it moves closer to the
drill string and thus a restoring force is apparent on the
anvil-bit. It can be called an elastic restoring force because it
is generated as formation disintegration occurs during the bit
displacement and therefore has a rate of force buildup. This force
is a resistance force and is independent from the energy consumed
by damping and is actually the force that keeps the drill string
from falling through the formation, yet in the operating system it
can be called an elastic restoring force of the formation. Yet
another restoring force is the mechanical action of the bit that
tends to move the anvil-bit toward the drill string as it rotates
and moves the bit cutting structure onto new formation. The elastic
restoring force of the formation then may be potential energy
returned from the formation, formation resistance to the bit load
force, and forces due to the mechanical action of the rotating
bit.
Due to gravitational effect or other forces, the drill string rests
on the anvil thrust ring springs which in turn biases the anvil-bit
which in turn exerts forces on the formation. These forces on the
formation cause the portion immediately adjacent the bit to exhibit
visco-elastic characteristics somewhat like a damped spring,
according to its elasticity, its energy absorbtion rate and other
factors. In this static situation with the anvil-bit elastically
suspended., the position at which the anvil-bit mass center rests
undisturbed will be considered its equilibrium position, X.sub.o.
X.sub.m will denote the amplitude or maximum displacement of the
mass center from equilibrium. Although various other movements will
take place during system operation, such as drill string rotation
and its downward progression, only the theoretical aspects of the
dynamic reciprocal steady-state motions will be considered. The
direction of drill string progression which would normally be
downward will be considered the positive direction and the opposite
direction will be the negative direction. Damping as used
throughout this specification is energy dissipation that would tend
to diminish the amplitude of oscillations in the vibrational
system, and although some energy reduction will be from ordinary
operating machine losses most damping will occur as energy usage in
formation disintegration.
The hammer mass m.sub.1 provides the periodic impulsive exciting
inertial forcing function F(t) for anvil-bit mass excitation and is
a part of the reciprocal fluid motor. m.sub.2 is the anvil-bit
mass, has a fixed value (weight) and is a part of the spring mass
oscillatory sytem that reciprocates longitudinally under the
influence of the elastic restoring forces K, the damping
coefficient C, the percussive blow F(t) of the hammer, and the
force W applied by the drill string. m.sub.3 is the mass of the
drill string or other equivalent force means that provide the force
W exerted on the spring mass system. These masses and forces are
shown in FIG. 17, and determine in a large part the vibrational
system performance. The anvil-bit mass, m.sub.2, will behave
according to the fundamental formula, F=ma and modified for the
specific conditions, the formula describing the motion of the
center of mass, m.sub.2, is F (t)+W = m.sub.2 X + C.sub.1 X +
C.sub.2 X + K.sub.1 X + K.sub.2 X where: F is the percussive blow
force, (t) is the transfer time of the blow force, and F (t) may be
called the system forcing function. W is the drill collar weight
force applied to the anvil-bit. X is acceleration in the X
direction and X is a velocity in the X direction. C.sub.1 is the
damping coefficient of the anvil thrust rings, and C.sub.2 is the
damping coefficient of the formation adjacent the bit and combined
make up the total damping coefficient, C. K.sub.1 is the spring
constant of the anvil thrust rings and K.sub.2 is the spring
constant of the formation adjacent the bit and combined make up the
total spring constant, K. X is the displacement of the center of
mass from the equilibrium position, X.sub.o. FIG. 17 shows these
factors affecting the anvil-bit mass. The use of the above equation
allows predictable behavior of the anvil-bit mass and is the basis
underlying this portion of system operation. Two desirable and
distinct types of steady state vibrational system operation or
modes can be achieved with the percussion tool. The mode is
dependent on the exciting, elastic and damping characteristics of
the system and will be described with specific limits but without
assigning actual values to the controlling factors. An important
aspect of each operational mode is the phase angle relationship of
the power inputs to the receiprocating anvil-bit mass m.sub.2. The
driving or exciting motor mass m.sub.1 percussive blow F(t) is
approximately 180.degree. from the driven anvil bit mass m.sub.2 in
the stiffness controlled system operational mode while in the mass
controlled system operational mode, m.sub.1 leads m.sub.2 by
90.degree. with an allowable variation of plus or minus 90.degree..
In addition to these primary power input phase angles with the
percussive motor, the secondary power input phase angle is also of
major significance since it may add or detract from the primary
power input to the anvil-bit mass. The secondary power input
resulting from the drill collar weight force W coincides with the
downward anvil-bit travel as the anvil thrust rings expand and
release their energy. This is to say that the energy from F(t) and
W added to the reciprocating m.sub.2 is transferred and consumed
inphase and thus these two power inputs compliment each other. This
power consumption occurs from 0.degree. to 180.degree. in the
stiffness controlled system and from -90.degree. to +90.degree. in
the mass controlled system. Both power sources then add energy to
the anvil-bit mass m.sub.2 from top dead center to bottom dead
center in the positive downward travel of the anvil-bit.
The first operational mode is the non-resonant frequency mode,
M.sub.1, and would normally be used in drilling the softer or more
brittle formations where maximum energy for formation fracture
would not be required and where a greater part of all energy would
be directly used in formation disintegration. This operational mode
is a critically or overdamped, stiffness controlled, forced,
vibrational system in which the directly applied forces of drill
collar weight, W, and percussive blow, F(t), are of primary
importance while the mass of the anvil-bit is a comparatively
insignificant factor.
In operation the anvil-bit mass, m.sub.2, is repeatedly displaced
downardly, a distance, X.sub.m, relating to the applied impulsive
force, F(t) and force, W, and returns to the equilibrium position,
X.sub.o, in a controlled manner due to a particular type of damping
in the elastic restoring forces acting on it and specifically may
be critically damped or over damped. This is to say that the
combined restoring forces, comprised of the formation and the anvil
thrust ring spring, returns the anvil-bit to its equilibrium
position without causing it to "overshoot" or pass the equilibrium
position between the percussive impulses. In the case of critical
damping the return is done in the shortest possible time with the
mass involved but in the overdamped condition the return time may
be longer. The limit on this non-resonant operational mode damping
factor C/C.sub.c is equal to or greater than one (1), where the
damping factor C/C.sub.c is the ratio of actual combined damping to
system critical damping, C.sub.c, which is .sqroot.4Km.sub.2. This
simply stated is that the combined damping is equal to
.sqroot.4Km.sub.2 or greater. FIG. 18 graphically pictures the
motions described by the center of mass in the criticaly damped
condition by the curve designated C.sub.c and in the overdamped
condition by the curve designated C.sub.o. The motion described by
the center of mass of the anvil-bit theoretically does not pass
across the equilibrium position but only returns to it. The
movements of the anvil-bit under actual conditions of operation may
be very complex and the descriptions herein are not meant to
emcompass the various minor movements caused by the rolling,
climbing action of the bit or other factors but to generally
describe the predominant characteristics of the damping factors and
the elastic restoring forces as effects the anvil-bit.
Considering that a cycle starts at percussion which would be
0.degree. and ends again at 0.degree. or 360.degree., the next
successive percussive blow, the maximum displacement, X.sub.m,
would be at 180.degree. and from this point through the remainder
of the cycle, the center of mass of m.sub.2 returns to or
approaches X.sub.o and does so exponentically or asymptotically
with time depending on the particular degree of damping; but as
before stated is .sqroot.4Km.sub.2 or greater. Then depending on
the spacing of periodic impulses of the forcing function F(t), the
center of mass of m.sub.2 is at or near X.sub.o at cycle start and
end. Then it can be stated that weight or force energy, W, that has
been stored over the time period, S, (FIG. 13) and used in doing
work during time period, Y, is in phase with the primary power
impulsive force, F, and will compliment forcing function energy
usage. The fact that the center of mass of m.sub.2 may not have
reached X.sub.o between the impulse forces in the overdamped
condition is of little consequence since m.sub.2 has reached the
major part of its negative travel to X.sub.o and in actuality
X.sub.o will adjust downwardly to the center of mass of m.sub.2.
The important point is that the energy usage from the forcing
function, F(t), and the force energy, W, coincide and are used in
phase. This is a considerable departure from the standard industry
conditions where the percussive blow occurs, negates the drill
collar weight, and then the percussive blow work is done without
the assistance of the drill collar weight energy. Then somewhere
about 180.degree. out of phase with the percussive blow, the drill
collar weight energy work is done and used without the assistance
of the percussive blow tending to provide an uncoordinated system
with little or no advantage over the conventional drilling system
except in special cases where suitable drill collar weight or force
cannot be applied to the system.
When the anvil-bit is driven downwardly on each cycle from its
equilibrium position by the percussive blow, F(t), the
displacement, X.sub.m, and force are positive and will be exerted
on the formation in addition to the drill collar load forces, W,
also exerted positively downardly on the formation. These forces F
and W have like signs and are additive indicating superposition
force. The following formula would express the mode forces exerted
on the formation in operation of the system. F.sub.M1 = F + W.
The second operational mode, M.sub.2, is resonant frequency system
operation. This harmonic, underdamped vibrational system is said to
be mass controlled and involves exciting the anvil-bit mass to a
frequency substantially equal to its natural or fundamental
frequency and thereby not only producing a percussive blow force
superpositioned on drill collar weight force but also an additional
force magnification, and maximum force transmissibility phenomena.
Generation of this phenomena can be a worthwhile contribution to
the effectivity of the system and would normally be applied in
drilling the harder formations where the extra force output is
needed and the greater formation elasticity is encountered. In this
case the system operation can better be explained through
successive tool cycles because the reinforcement phenomena occurs
over a continuous series of cycles.
The periodic impulsive hammer percussive blow excites or displaces
the anvil-bit mass downwardly from the equilibrium position
imparting energy to it in the form of velocity. The anvil-bit then
has momentum because of its mass and velocity and will produce
formation forces relating to these factors. A portion of this
kinetic energy will be used to "drill" formation and this is a form
of damping designated as C.sub.2. The remaining energy will be
returned from the formation to the anvil-bit because of formation
elasticity or spring rate, K.sub.2, in the form of veloicty or
rebound in the opposite or negative X direction. The inertial
effect of the anvil-bit mass velocity causes it to "overshoot" the
equilibrium position and causes anvil thrust ring spring
compression, potential energy storage in them in addition to static
equilibrium force, W, energy until the mass again slows down,
reverses direction and accelerates in the positive direction. The
elastic restoring forces in the formation and the anvil thrust ring
springs always tend to return the anvil-bit mass center to
equilibrium while the percussive blow continually disturbs this
tendency. As the anvil-bit moves downwardly again another inphase
positive impulse force, F(t), is added to the positive anvil-bit
motion, thereby adding more energy to the reciprocating mass,
m.sub.2, and reinforces the movement. The forces that now can be
exerted on the formation are the static weight drill collar forces,
percussive forces imparted to the anvil-bit by the hammer, and the
forces supplied by the inertial kinetic energy in the anvil-bit and
the potential energy of the anvil thrust ring springs. On each
cycle as above described, the anvil-bit forces are such as to cause
it to pass the equilibrium position and the system will be storing
energy that is not used drilling the formation and over a series of
cycles can develop steady-state forces that can be said to be
magnified.
The anvil-bit has a natural undamped frequency, p.sub.u, equal to
.sqroot.K/m.sub.2, and a modification of this formula with the
damping factor, n = C/2m.sub.2, yeilds the system anvil-bit natural
damped frequency p = .sqroot.p.sub.U.sup. 2 - n.sup.2. When the
anvil-bit mass m.sub.2, is excited by the forced frequency w,
hammer forcing function F(t), to equal its natural damped
frequency, resonance occurs and the inphase impulse force
conditions are met for force reinforcement and force magnification
as described above. This is to point out that a first relationship
between the anvil-bit mass m.sub.2 and the combined system
restoring forces K is such that the forces K must deliver the mass
m.sub.2 to certain advantageous positions within specified limits
during steady-state oscillation, and a second relationship between
the anvil-bit mass m.sub.2 and the reciprocating hammer mass
m.sub.1 is such that percussion of the masses m.sub.1 and m.sub.2
must occur within said certain advantageous position limits.
Theoretically, if the forced frequency w were equal to the undamped
natural frequency P.sub.u, the magnification factor and amplitude
would be very large if no physical restraints were placed on the
system in the form of damping. This can be seen from FIG. 19. The
system, however, has a fairly high range of damping but always less
than .sqroot.4Km.sub.2 that is said to be variable because of the
formation variable, C.sub.2. This damping of the oscillating
anvil-bit occurs mostly as energy usage in drilling the formation
which may vary significantly and would contribute the major part of
the damping coefficient, C, made up of C.sub.1 and C.sub.2. This
variable, C.sub.2, also will determine the numerical value of the
ratio of actual damping, C, to the critical damping, C.sub.c. The
damping factor C/C.sub.c may range from near zero (0) for the hard
elastic formations to unity (1) for the softer energy absorbing
formations and because of this variation it is one of the major
determinants of the actual values of the magnification factor,
X.sub. m /(F(t contribution to )/K). Another major magnification
factor is the ratio of forced frequency to natural frequency. The
effects associated with resonance are spread over a broad range of
frequencies so specifically percussion should occur anytime during
the positive travel of the anvil-bit mass. This means that from
resonance the variation of forced frequency to natural damped
frequency may arbitrarily range from no less than 0.5/1 to no more
than .sqroot.2/1 and within these limits magnification occurs.
The formula ##SPC1## expresses the magnification factor in terms of
the frequency ratio, w/p.sub.u, and the damping factor C/C.sub.c.
It may be used to determine the amplitude, A, of the steady state
vibration produced by the impressed force, F(t). This relationship
of magnification factor, the frequency ratio, and the damping
factor is also shown graphically in FIG. 19. Thus it can be seen
that considerable force magnification can be generated at or near
system resonance in tough formations.
Considering the foregoing, the following formula would
approximately express the system produced formation forces when
operating in mode, M.sub.2. ##SPC2## showing that the forces
exerted on the formation would be drill collar load forces plus a
magnified percussive blow force,
As previously stated this tool is capable of two operational modes
and both have been described herebefore. The tool design paramaters
may be set so the system would always be a critically or overdamped
stiffness controlled system and therefore would be assured of
inphase superposition force operation. But each operational mode
has a broad frequency range and may be made to overlap. Then it is
also possible to provide design paramaters that would allow the
tool to change operational modes when it encounters significant
changes in formation characteristics. That is, if the tool design
were set for some arbitrary condition of critical damping, any
actual formation damping change could cause the damping to switch
from the critical damping condition and allow mode change. Of cours
when mode M.sub.2 conditions existed, the forced frequency would
have to substantially match the anvil-bit natural frequency. This
could be easily adjusted by changing the fluid supply pressure if
required, but the natural feed-back couple inherent in a
conservative intertial interchange system would somewhat
self-adjust or tune itself according to the hardness of the
formation encountered and therefore adjust cycle frequency
accordingly. Also, the resonant effects are spread over a broad
variation range and would further tend to make it comparatively
easy to operate the system within the specified limits described by
the 0.5:1 and the .sqroot.2 :1 ratios. It is an object to maintain
broad operating resonant frequency ranges in the mass controlled
system so that the whole spectrum of tool operation can be covered
without underdamped non-resonant tool operation. This specification
sets specific limits on the invention with damping factors of one
(unity) or more for operation in the non-resonant frequency mode
stiffness controlled system and with damping factors of one (unity)
or less for operation in the resonant frequency mode with 0.5/1 and
.sqroot.2/1 forced frequency to natural frequency ratio resonance
limits in a mass controlled system.
No other known device of this type used for the same purposes
exhibits the characteristics of this tool. The tools in use conform
to standard industry practice and do not coodinate in like manner
energy usage from the weight or force applied on the percussion
drill and do not even remotely suggest the disclosures herein.
Also, then this specification sets further limits that energy usage
for formation disintegration of all operations in either mode be
generally inpphase, coincide or substantially overlap.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a transverse, vertical, sectional view of the upper
portion of the percussion tool B showing it connected to the drill
collar A. The hammer location is shown at percussion and the
central valve element shifted relatively downward as it would be
momentarily after percussion and valves are positioned for piston
upstroke;
FIG. 2 is a continuation of FIG. 1 showing the lower portion of the
percussion tool with a conventional rotary drill bit C attached at
its lowermost end;
FIG. 3 is a reduced diagrammatic longitudinal sectional view
illustrating fluid flow to below the hammer from the drill stem and
from above the hammer into the central anvil bore while piston is
moving upwardly;
FIG. 4 is a reduced diagrammatic view similar to FIG. 3
illustrating fluid flow into the chamber above the hammer and
exhaust fluid flow from the chamber below the hammer while the
hammer is on the downstroke;
FIG. 5 is a reduced diagrammatic longitudinal sectional view of the
lower portion of the percussion tool illustrating anvil extension
distance X at percussion and the reactionary period thereafter
caused by a combination of percussive force F and anvil thrust ring
force W;
FIG. 6 is a horizontal, cross-sectional view taken on line 6--6 of
FIG. 1;
FIG. 7 is a horizontal, cross-sectional view taken on line 7--7 of
FIG. 1;
FIG. 8 is a horizontal, cross-sectional view taken on line 8--8 of
FIG. 2;
FIG. 9 is a fragmentary view similar to a portion of FIG. 2 except
illustrating an anvil thrust ring spring of an alternate
construction;
FIG. 10 is a longitudinal cut-away view of the percussion tool in
the area of the valves showing the components as in a portion of
FIG. 1 except that cut-away sectional view taken is rotated
45.degree. from that in FIG. 1 and shows the vertical passageway
through the valve element as well as a means of support for valve
seat members;
FIG. 11 is a perspective view of one design configuration of an
anvil thrust ring;
FIG. 12 is a traverse vertical sectional view similar to a portion
of FIG. 1 in the area of the lower end of top sub illustrating an
alternate construction that embodies spring bumper rings on the top
sub;
FIG. 13 is a graph in which the tool applied forces on the
anvil-bit are plotted with respect to time during operation of the
percussion tool and this graph shows the inter-relationship of
these paramaters in a typical manner;
FIG. 14 is a view similar to FIG. 9 showing an anvil thrust ring
spring of different design;
FIG. 15 is an enlarged perspective cut-away view of the valve 4
showing the passages and ports in it, its four valve faces, and its
overall configuration;
FIG. 16 is an enlarged cut-away perspective view of adapters 34 and
37, valve seat ring 33, and locator ring 52 as they are positioned
relative to each in assembly;
FIG. 17 is a schematic model diagram of the springmass vibrational
system of the percussion tool and graphically represents the
theoretical concepts involved in system function, the forces
involved in system operation, and other factors involved in the
invention;
FIG. 18 is a graphical comparison of the motion described by the
center of mass, m.sub.2, of the anvil-bit during steady state
operation when the system is overdamped, critically damped, and
underdamped;
FIG. 19 is a graph in which the amplitude or magnification factor
has been plotted against the frequency ratio for the various values
of the damping factor during resonant frequency or mode M.sub.2
operation;
FIG. 20 is a longitudinal quarter sectional cutaway view in the
area of the valves showing an alternate construction of the valve
cage and valve element;
FIG. 21 is a transverse, vertical view of the upper portion of the
percussion tool B' with a quarter sectional transverse cut-away
view;
FIG. 22 is a continuation of FIG. 21 showing the lower portion of
the percussion tool with a conventional solid button BUTTON bit,
C', as an integral part of the anvil;
FIG. 23 is a longitudinal quarter sectional cutaway view similar to
the view in FIG. 21 in the area of the valve cage but rotated so
that a section is taken through the set of alternate passageways
through the valve cage; and
FIG. 24 is a horizontal, cross-sectional view taken on line 24--24
of FIG. 21.
DETAILED DESCRIPTION
Structure to carry out the objects set forth and the basic
functional performance of two mode operation, superposition force,
inphase energy consumption, force magnification, and formation
cooperation with the drilling means are described hereafter. These
constructions are for the most part the same but will vary to some
extent because of existing equipment with which they must function,
because of historical practices that came into being over the
years, and because of difference in application which varies with
the industry involved. While the oil and gas industry mostly employ
liquids as the system operating fluid medium, the mining, blast
hole, and construction industry employ mostly a gaseous fluid,
primarily air, as the system circulating medium. The device must be
capable of operation on a variety of fluids but fluids vary
considerably in compressibility, viscosity and physical state such
as gaseous or liquid. Therefore, although the basic ideas and
constructions involved are primarily the same, a variation in
valving means is included to provide optimum performance and
optimum economy according to the fluid used in the system.
Constructions shown in FIGS. 1, 2, 10, and 20 are specifically
provided for relatively non-compressible fluids but could be used
on compressible fluids while constructions shown in FIGS. 21, 22,
and 23 are specifically for compressible fluids and will not
operate on relatively non-compressible fluids. This alternate
embodiment is limited to the valving means for hammer oscillation
to provide tool simplicity and manufacturing economy when the tool
is operated from a gaseous pressure source.
With more detailed reference to the drawings, the letter A
designates the lowermost end of the drill collars in a string of
rotary drill pipe which supplies the torque for turning the bit,
weight loadings for the bit and system circulating fluid and on
which a percussion drilling tool B is mounted. A rotary drill bit
C, which may be of conventional tooth or button type, threadedly
engages the lower end of the rotary percussion tool for drilling a
borehole. Although a threaded detachable bit is shown, the bit may
be made an integral part of the anvil and be of solid fixed design
as shown in FIG. 22.
The percussion tool B generally is made up of the housing 1, anvil
2, hammer 3, valve 4, and anvil thrust rings 5, each of which will
be described in detail.
The housing 1 is comprised of the top sub 6, barrel 7, and driver
sub 8 and serves, among other things, as a casing for the hammer,
guide for the anvil and a means of attachment to the lower end of
the drill string as well as a means of fluid containment and
conductance. The top sub 6 threadedly engages the lower end of the
drill collars A and has a lower face 9 and an axial bore 10 for
passage of drilling fluid used to power the percussion tool. The
lower end of top sub 6 is screwthreaded, as indicated at 12, to
threadedly engage the upper end of tubular barrel 7. The lower end
of top sub 6 has packing elements 13 surrounding the bore 10
thereof to form a fluid-tight seal with the upper tubular end 14 of
valve 4 extending into bore 10. The lower end face 9 may have a
plurality of radial grooves to allow the unrestricted pressure
fluid flow in chamber formed above hammer. A design variation of
lower end face 9 is shown in FIG. 12 and may be a desirable feature
in some applications. The configuration incorporates a spring
element 59 mounted on a lower end face 9 for coaction with upper
end face 28, of piston 25 for reversing hammer 3 at the termination
of its upstroke.
The barrel 7 is a cylindrical casing threadedly attached to top sub
6 on its upper end and to the driver sub 8 on its lower end, as
indicated at 22 and surrounds the hammer 3 and upper end of anvil
2.
The driver sub 8 is generally tubular in shape and is threadedly
attached to the lower end of barrel 7 and forms the lower end of
housing 1. The driver sub 8 has a lower face 24 and an internal
bore that forms a bushing guide for the anvil 2. The driver sub 8
also has a plurality of internal grooves formed longitudinally
thereof, as indicated at 15, FIGS. 2 and 8, to receive pins 16 in
sliding relation thereto.
The anvil 2 is slidably connected to housing 1 and rotatable
therewith and extends through driver sub 8 and into lower end of
barrel 7 and on its lowermost end may threadedly engage drill bit
C. The anvil 2 has an upper end face 17 that may have a plurality
of radial grooves to facilitate fluid flow between said face and
oppositely facing hammer face. An axial bore 18 is formed centrally
of anvil 2 for passage of drilling fluid from the lower tubular end
20 of valve 4 to the bit. The upper portion of bore 18 has packing
elements 19 surrounding the bore 18 thereof to form a fluid-tight
seal with the lower tubular end 20 of valve 4 extending into bore
18.
The anvil 2 has a plurality of recesses 21 on its shank
complimentary to grooves 15 of driver sub 8 to receive a portion of
the pins 16, the arrangement being such that a splined joint is
formed allowing longitudinal movement of the anvil 2 relative to
the housing 1 but no appreciable rotational movement between said
members. The shank of anvil 2 near the upper end is grooved to
receive a split annular retainer or snap ring 61 to assemble the
anvil with the driver sub 8 and provides a means to limit the
longitudinal movement between said anvil 2 and driver sub 8.
Packings 26 are interposed in fluid-tight sealing arrangement
between driver sub 8 and anvil 2 to prevent fluid flow
therebetween. Anvil 2 intermediate its ends has a shoulder 23 that
together with lower face 24 of driver sub 8 and shank of anvil
confine anvil thrust rings 5.
The anvil 2 may be a separate part from the bit as shown in FIG. 2
or may be an integral part with the bit as shown in FIG. 22 but in
either case the anvil and bit together form the mass, m.sub.2, of
the spring mass system shown in FIG. 17. The anvil 2 along with the
bit is suspended by the elastic restoring forces of the system and
is adapted to oscillate axially while the drilling tool
rotates.
Anvil thrust rings 5 are extensive mechanical force storage and
energy usage phase transfer means in the form of annular spring
elements of various configurations and materials that in usage
provide a resilient bias between shoulder 23 of anvil 2 and lower
face 24 of driver sub 8. The anvil thrust rings may be thick, high
force, low inertia, short reaction time, frustro-conical annular
springs as indicated in FIGS. 2, 11, and 22 or radial ring spring
shape as indicated in FIG. 14 by 5A. Another anvil thrust ring
spring design is indicated in FIG. 9 by 5B in which flat annular
metal rings are alternately stacked with an elastomer and having
bonded interfaces. The function of the anvil thrust rings is to
store various forces exerted on them and to return said forces to
the confining shoulders. Specifically they store and transfer force
energy exerted by the tool housing 1 to the anvil 2 and as
indicated in FIGS. 5, 13, and 17 by W. They also extend
simultaneously with anvil 2 downward movement due to the percussive
blow F(t) and coact with said blow to simultaneously release the
energy stored therein for superposition bit forces on the formation
as used in mode M.sub.1 tool operation. In this function and mode
they also change the phase of drill collar weight energy
consumption about 180.degree., depending on conditions, and puts
said energy consumption in phase with the percussive blow for more
effective energy usage. FIG. 13 graphically shows the phase
alignment and hense superposition force. In addition to the above
the anvil thrust rings 5 in mode M.sub.2 operation provide the tool
portion of the elastic restoring forces K and the damping
coefficient C to the oscillating anvil-bit. This elastic restoring
force is indicated in FIG. 17 by K.sub.1 and the damping
coefficient is indicated by C.sub.1.
The formation, O, being drilled plays a large role in system
operation as it cooperates with the system to provide damping, the
energy consumption of formation disintegration, and a portion of
the elastic restoring forces, the energy returned to the system by
the formation. The formation damping coefficient is indicated in
FIG. 17 by C.sub.2 and the formation elastic restoring force is
indicated by K.sub.2. The formation then consumes system energy or
returns it to the system in varying degrees depending on the makeup
of the particular formation.
Hammer 3 consists primarily of piston 25, adapters 34 and 37, valve
seat ring 33 and locator ring 52. Piston 25, adapters 34 and 37,
and valve seat ring 33 also form a valve cage internally in hammer
3 intermediate its ends and as defined by the chamber formed
between said parts and enlarged portion of valve 4.
Piston 25 of hammer 3 is reciprocally mounted in barrel 7 of
housing 1 for fluid biased reciprocation therein for imparting
percussive blows to the anvil 2. Outside diameter of piston 25 is
in fluid-tight sealing arrangement with bore of barrel 7 and has a
lower end face 27 oppositely facing to upper end face 17 of anvil 2
and adapted for coaction therewith. Piston 25 also has an upper end
face 28 facing opposite to end face 9 of top sub 6 and said faces
may be used for coaction. Piston 25 has a stepped internal bore
extending centrally from end to end and is used for fluid passages
and mountings for the various parts forming the valve cage.
The valve cage is formed on its top side by adapter 34 which has a
valve seat 30, a central bore surrounding upper tubular end 14 of
valve 4 and is in sliding relation thereto and in sealing
arrangement therewith by seals 35, and has a concentrically spaced
passage formed in part by its outer sides and radially divergent
ports through its downwardly depending skirt to allow fluid flow
between chamber 36 above hammer and valve cage. Adapter 34 may be
screwthreaded in piston 25 or otherwise rigidly mounted thereto and
serves to rigidly fix locator ring 52 and adapter 37 in enlarged
bore of piston 25.
Adapter 37 has a valve seat 32, a central bore surrounding enlarged
mating surfaces of valve 4 and is in sliding arrangement thereto
and in fluid-tight sealing arrangement therewith by seals 38,
concentrically spaced passages formed in part on its outer surface,
an upwardly depending ported skirt for spaced relation with adapter
34 and locator ring 52, and downwardly extending legs for spaced
relation with reduced piston 25 bore diameter. The adapter 37
skirt, ports, passages and legs being such as to allow fluid
communication between valve cage, passage 42 between piston bore 44
and outer surface of lower tubular end 20 of valve 4 and chamber 39
below the hammer.
Locator ring 52 may be rectangular in cross section and annular in
shape and split one place to allow expansion assembly in external
groove of valve seat ring 33 and is rigidly fixed in location by
upwardly depending ported skirt of adapter 37 and the downwardly
depending ported skirt of adapter 34. Locator ring 52 serves to
limit the axial sliding movement of valve seat ring 33 relative to
adapters 34 and 37.
Annular valve seat ring 33 has two seat surfaces 29 and 31 which
may be frustro-conical in shape and has an outside centermost
groove complimentary to but wider than locator ring 52 for limited
sliding arrangement therewith. Outside surface of valve seat ring
33 is mounted for axial limited sliding movement in skirts of
adapters 34 and 37 and is in fluid-tight sealing arrangement
therewith by seals 60, provided in grooves in the outside
diameter.
Valve 4 is a single element that conducts pressure fluid to and
from the valve cage and controls charging and discharging of
chamber 36 above hammer 3 and chamber 39 below hammer 3, thus
creating differential pressures alternately across the hammer
thereby causing reciprocation of same. Valve 4 which may be made of
two or more joined pieces and screwthreaded or otherwise fixed
together is mounted in hammer 3 for limited longitudinal
reciprocation therewith and is adapted to shift relative to hammer
at termination of each up and down hammer stroke to redirect
pressure fluid. Valve 4 has an upper tubular end 14 extending
through adapter 34 and into bore 10 of top sub 6 and a lower
tubular end 20 extending downwardly through adapter 37, bore 44 of
piston 25 and into upper end of bore 18 of anvil 2. Valve 4 has an
enlarged spool-like portion which joins the upper and lower tubular
ends and has four valve faces 45, 46, 47, and 48, FIGS. 1 and 15,
formed to generally compliment internal shape of valve cage and are
in fixed spaced relation to each other. Valve faces 45, 46, 47, and
48 mate and coact with valve seats 30, 29, 31, and 32, FIGS. 1 and
16, but the spacing of said valve seats and valve faces are such
that when valve 4 is shifted downwardly relative to hammer 3 valve
face 48 and valve seat 32 are closed and valve face 46 and valve
seat 29 are closed. Conversely, when valve 4 is shifted upwardly
with respect to hammer 3, valve face 45 is closed with valve seat
30 and valve face 47 is closed with valve seat 31. The valve face
and seat arrangement is such that when two pressure fluid passages
between a valve face and its respective valve seat are closed, two
pressure fluid passages between a valve face and its respective
valve seat are open.
Valve 4 has a restrictive fluid passageway 49 communicating with
bore of upper tubular end 14 and lower tubular end 20 to provide
bypass fluid through the tool. Passageway 49 may be designed to be
easily replaceable for size adjustment to accommodate various flows
as may be required for well conditions and associated equipment
used, but must be restrictive to create a differential fluid
pressure between bore of upper tubular end and lower tubular end of
valve 4 for the percussion tool to operate.
Upper tubular end 14 of valve 4 above restrictive passageway 49 has
a multiplicity of radially divergent passages 50 allowing pressure
fluid flow from bore of upper tubular end 14 to valve cage space
between valve faces 46 and 47 of valve 4. Lower tubular end 20 of
valve 4 below passageway 49 has a multiplicity of radially
divergent passages 51 allowing fluid communication with bore of
lower tubular end 20 and valve cage space inside adapter 37.
Enlarged portion of valve 4 also has a multiplicity of
concentrically spaced passages 40, FIGS. 10, 15 and 6,
longitudinally therethrough that allows fluid communication with
lower tubular end 20 and valve cage space formed by a recess inside
valve face 45.
Now it is apparent when valve 4 is shifted downwardly relative to
hammer 3 that pressure fluid can flow from bore 10 through upper
tubular end 14 and passages 50, between valve face 47 and valve
seat 31, through ports and passage of adapter 37, through passage
42 and into chamber 39 below hammer to apply pressure force to act
on underside of hammer 3 and, simultaneously, exhaust fluid can
flow from chamber 36 above hammer 3 through passages formed in part
by adapter 34, between valve face 45 and valve seat 30, through
passages 40 and lower tubular end 20 allowing fluid pressure
dissipation above hammer and acceleration of same upwardly. This is
shown diagrammatically in FIG. 3. It is also apparent that when
valve 4 is shifted upwardly relative to hammer 3 that pressure
fluid can flow from bore 10 through upper tubular end 14 and
passages 50, between valve face 46 and valve seat 29, through ports
and passages of adapter 34 into chamber 36 above hammer 3 to exert
fluid pressure thereon and exhaust fluid can flow from chamber 39
below hammer 3 through passage 42 around and through passages
formed in part by adapter 37, between open valve face 48 and valve
seat 32, through passages 51 and lower tubular end 20 of valve 4
allowing fluid pressure dissipation below hammer 3 and downward
acceleration of same due to differential pressure across it. This
is shown diagrammatically in FIG. 4. Thus, it can be seen that two
valve faces 46 and 47 control and direct the higher pressure fluid
to chambers 36 and 39 while two valve faces 45 and 48 control and
direct the lower pressure exhaust fluid from chambers 36 and 39 and
that a change in position of valve 4 which carries said valve faces
would switch differential fluid pressures across hammer 3.
Valve 4 has upwardly facing surfaces 53 and 54 constantly exposed
to supply pressure fluid and downwardly facing surfaces 55 and 56
constantly exposed to exhaust fluid and since supply fluid pressure
is always greater than exhaust fluid pressure, valve 4 has a
continual downward bias from fluid acting on said surfaces. Valve 4
also has a generally downward facing surface 57 constantly exposed
to pressure fluid within passage 42 and chamber 39 below piston and
an upwardly facing surface area 62 equal to area of surface 57
constantly exposed to pressure fluid in lower tubular end 20 of
valve 4. When supply pressure fluid is valved to chamber 39 below
hammer 3, fluid acts on surface 57 and opposes downward bias caused
by pressure acting on surfaces 53 and 54, and surface 57 may be
sized to provide a net bias on valve 4 in the upward direction when
pressure fluid is acting on lower surface 27 of hammer 3 to raise
it. Valve 4 then is biased upwardly when hammer 3 is on its
upstroke and biased downwardly when hammer is on its downstroke.
Pressure fluid on other surfaces of valve 4 may be balanced
vertically and radially or unbalanced in the vertical direction to
some degree to aid valve shift. This can be done by varying areas
exposed to pressure fluid on spool-like flanges of valve 4 and seat
ring 33 which is fluid biased in the direction that would tend to
unseat closed exhaust valve faces 45 or 48. Although valve 4 is
biased in the direction of hammer 3 travel, inertia holds the valve
positioned to continue feeding pressure fluid for acceleration of
said hammer movement until hammer is rapidly decelerated as when it
strikes the anvil 2 or top sub 6 at which time the inertia of valve
4 and the pressure bias in the direction of travel will shift the
valve very rapidly, and precisely at termination of hammer stroke.
It is, therefore, important to maintain certain relations of
weight, bias, and acceleration of valve 4 to hammer 3. Valve
operation is "on demand" and is responsive and cooperative to
hammer accelerations and velocities.
It should be noted here that by proper sizing of fluid bias area
acting on valve 4 relative to mass of hammer 3 and speed of same,
this tool can be made operational without the hammer striking an
upper internal end face such as face 9 of top sub 6 as in some
cases may be desirable. This can be done by providing an upward
bias on valve 4 in the hammer 3 upstroke that will cause said valve
to accelerate upwardly faster than the hammer acceleration at some
point in the upward travel of said hammer such as the hammer 3
leveling off to a constant velocity and a zero rate of acceleration
due to limited pressure fluid supply. This arrangement would allow
hammer upstroke length variation to be determined by supply fluid
pressure, terminal velocity of hammer on its downstroke and
relative masses of hammer and anvil as well as hardness of the
formation being drilled and other factors. Of course, in any given
device with constant fluid supply pressure, the only appreciable
variable would be the formation being drilled and the tool would
tune itself by stroke length adjustment and associated frequency,
energy output, etc., to the characteristics of the formation.
Lower tubular end 20 of the valve 4 may have one or more ports 58
through its wall constantly communicating with the bore of lower
tubular end 20 and passage 42 for fluid passageway therethrough.
These ports can provide a means for controlling the hammer upstroke
velocity by reducing the fluid pressure applied to raise the
hammer. They also provide better fluid dissipation below hammer on
its downstroke.
A VALVE AND VALVE CAGE ALTERNATE CONSTRUCTION IS SHOWN IN FIG. 20.
This design functions in the same manner as the valve shown in FIG.
1 but is of a simpler configuration made possible with the use of
better erosion resistant materials and metalurgy. Also a spring has
been incorporated that biases the valve upward and allows tool to
begin cycling easier. This spring force aids initial valve shift
when valve inertial forces are minimal.
An enlarged spool-like portion of valve 4 has external cylindrical
and oppositely facing flat surfaces forming edges, 63 and 64, that
in cooperation with internal and end face surfaces 65 and 66 of
valve rings 67 control charge pressure fluid from passage 51.
Cylindrical surfaces of 63 and 64 are slightly smaller than
cylindrical surfaces of 65 and 66 of valve rings 67 and are
positioned concentrically. The construction is such that when the
fixed spaced edges 63 and 64 are shifted down relative to edges 65
and 66 pressure fluid can flow to the chamber 39 formed below
piston 25, and pressure fluid is substantially blocked to chamber
36, formed above piston 26. Conversely, when valve 4 is shifted
relatively up, coacting edges 63 and 64 move apart and allow charge
pressure fluid flow therebetween while edges 65 and 66 move
together and past each other essentially stopping the fluid flow
between said edges. Valve rings may be generally rectangular in
cross section and closely fitted to bore of piston 25 with a fluid
seal therebetween. Valve rings are spaced in bore of piston by thin
wall tubular spacers 68 and 69. Adapter 34 which is screwthreaded
into mating thread at top of piston clamps spacers 68 and 69, valve
rings 67 and adapter 37 in bore of piston 25. Spring 70 surrounds
upper portion of lower tubular end 20 of valve 4 and its ends are
nested against the enlarged surface 57 on the upper side, and
against hammer 3 on the lower side. The spring as installed is
somewhat compressed and maintains force between valve 4 and hammer
3.
The foregoing valve structures would be used for the most part on a
liquid fluid driven tool while the following constructions shown in
FIGS. 21, 22, 23 and 24 apply totally to a gaseous fluid driven
tool. The liquid and gaseous fluid circuits are the same and only
the circuit switching valve is of modified design. Also the housing
1, anvil 2, piston 25, and anvil thrust rings 5 are the same as
before described. The valving means, although similar in some
respects, has notable differences.
With more detailed reference to the Drawings, fluid switching valve
means are comprised of the valve 75, and the valve block or cage
76. Valve 75 is a single element that conducts pressure fluid to
and from the valve cage and in cooperation with the valve cage
controls charging and discharging of chamber 36 above hammer 3 and
chamber 39 below hammer 3. Hammer 3 in this embodiment is made up
of piston 25 and valve cage 76. Valve 75 has an upper tubular end
77, extending into bore 10 of top sub 6 with a seal 13
therebetween, a lower tubular end 78 extending into and may rest on
top portion of anvil 2 and in sealing arrangement therewith by seal
79. Although not shown, it is apparent that the valve 75 could
equally well be fastened to the top sub 6 since these parts are
stationary relative to each other and that this may be advantageous
due to the fact that the valve 75 would be free from the rapid
anvil 2 oscillations. Central portion 85 of valve 75 is solid or
may have a passage therethrough that blocks or substantially
restricts fluid flow from passage of upper tubular end 77 to
passage of lower tubular end 78. Upper tubular end of valve
conducts pressure fluid to the radially divergent side parts 80
above central restriction 85 while the lower tubular end conducts
fluid from the radially divergent side parts 81 into the bore of
anvil 2. Valve 75 is slidably mounted through central opening 82 of
valve cage 76, and outside surface of valve 75 is a close sliding
fit with opening 82 so that the surfaces in close proximity or
contact form a seal or greatly restricts fluid flow therebetween,
thus eliminating separate sealing means.
Valve cage 76 is a single element screwthreaded in piston 25 that
has a central bore opening 82, and two enlarged bores, 83 being the
upper bore and 84 the lower two diameter bore, concentric with
opening 82 and surrounding outside of valve 75. Valve cage also has
one or more passages 86 formed in part by valve cage 76 and piston
25 that allow fluid communication from space formed by bore 83 to
passage 42 and one or more passages 87 that allows fluid
communication from space formed by enlarged bore 84 to chamber 36
above hammer 3.
Now it can be seen from FIGS. 21 and 22 that when hammer is in its
lower position pressure fluid can flow down through upper tubular
end 77 of valve 75, through its side parts 80 into space formed by
bore 83, through passage 86 and 42 into chamber 39 to act on the
underside of the hammer to raise same. Similarly it can be seen
from FIGS. 21 and 23 that exhaust fluid can flow from chamber 36
through passage 87 and space formed by bores 84, through valve side
ports 81 and into lower tubular end 78 of valve 75 to allow fluid
pressure dissipation above hammer 3 and thus a differential
pressure across the hammer. Now when the hammer is in an up
position, and valving is switched to power the hammer downwardly,
the relative position of valve and cage parts have changed so that
pressure fluid will flow from the upper tubular end of valve 75,
through its side parts 80 and into space formed by the two diameter
bore 84, through passages 87 in valve cage 76 into chamber 36, thus
providing force on hammer to drive it downward. Similarly it can be
seen that since hammer is raised relative to valve ports 81 and
thus exposing said port below lower end of valve cage 76 that
exhaust fluid can flow from chamber 39 below hammer through passage
42, through exposed part 81 into lower tubular end 78 of valve 75
and thus allows fluid exhaustion and pressure dissipation below
hammer for creation of differential pressure across the hammer and
movement of same downward for the percussion blow and anvil
excitation. Since the valve is relatively stationary while the
hammer moves, sliding port valving action occurs before the end of
each up and down hammer stroke for alternate pressure fluid
application up to hammer and reciprocation of same. This type valve
action is permissible with a gaseous fluid and in fact highly
desirable since the time required to bring hammer actuation
pressure up to supply pressure is longer than when a liquid is used
and thus in effect allows early fluid valving to compensate for the
slower pressure buildup. Also the hammer is positively actuated in
the opposite direction at each end of its stroke, thus eliminating
"dead spots" at these positions. Fluid switching occurs
intermediate the hammer stroke ends and momentum forces of the
hammer are such as to carry out the valving action although the
actuation fluid flow is momentarily halted when the involved ports
do not allow fluid passage. The constant alteration of unbalanced
pressures across the hammer cause rapid oscillations of the hammer
and on each cycle imparts inertial forces to the anvil-bit
according to the magnitude of the pressure source, the specific
tool configuration and the formation being drilled. The object of
course is to increase the rate of penetration of the bit through
the formation and to accomplish this most expediently by exciting
the bit in a certain fashion.
FIG. 3 shows diagrammatically the charge pressure fluid flow from
the bore of the drill string into chamber 39 below hammer and the
simultaneous discharge pressure fluid flow from chamber 36 above
hammer into the anvil bore and the resultant hammer upstroke caused
by differential pressure across the hammer. FIG. 4 depicts
diagrammatically the charge pressure fluid flow from the bore of
the drill string into chamber 36 above hammer and the simultaneous
discharge of pressure fluid flow from chamber 39 below hammer into
the anvil bore and the direction of resultant hammer downstroke
caused by differential pressure across the hammer. During the
hammer upstroke and downstroke, the anvil thrust rings are also
releasing and then storing energy provided by drill collar weight
loadings and anvil-bit formation forces.
FIG. 5 illustrates the forces and occurrences that transpire at
percussion in mode M.sub.1. W indicates the drill collar load
applied to the bit by the anvil thrust rings before and at
percussion. F indicates the percussive force applied to the
anvil-bit. The combined forces of W and F cause anvil extension
relative to the housing and is indicated by X.
FIG. 13 further graphically explains the sequence and nature of
events before, during, and after percussion in a general and
typical manner in operational mode M.sub.1. On the graph, forces
would be zero at the bottom of graph and would increase upwardly
while time progresses from left to right. W indicates drill collar
load and is shown on the left part of the graph as before the
percussion tool is operating. Operation is indicated by the
variable wave curves and indicates forces that come into play
during operation and are plotted versus time. Z indicates the anvil
thrust ring energy cycle between percussive blows and consists of a
first part energy usage time Y and a second part energy storage
time S. During energy usage time Y the anvil thrust rings 5 coact
with the percussive blow to extend the anvil 2 relative to the
housing 1, thus expending all or part of their stored energy.
During energy storage time S the anvil thrust rings 5 absorb energy
because the drill collar load moves downwardly tending to compress
them. Drill collar load force applied to the bit as transferred by
the anvil thrust rings may range from a minimum as indicated by FO
between percussive blows to a maximum equal to drill collar weight
load force W at percussion. F indicates the force applied to the
anvil 2 and hence the bit by the percussive impact of the hammer on
the anvil and, as shown, is synchronized for application at maximum
drill collar load application W. Forces W and F coact and are added
arithmetically to produce a relatively high instantaneous total
load force spike, as indicated by FT, on the bit for formation
crushing or cleavage by forcibly extending the anvil some distance,
as indicated by X in FIG. 5, said force spike exceeding formation
resistance to contact areas of the bit. In the graph (tindicates
time percussive force is applied to the anvil-bit and V indicates
total tool cycle time. Bit loadings drop considerably after
expenditure of forces F and W and anvil extension, and then rise to
maximum again as the drill collars move downwardly and distort the
anvil thrust rings while the hammer progresses through its up and
down stroke to again deliver a percussive blow. FIG. 17 is a
schematic model of the percussion tool on the end of a drill string
in a borehole. The valving and pressure fluid have been omitted for
clarity. This diagram, a particular configuration of the classical
forced-damped-spring-mass system, indicates the factors affecting
the anvil-bit, controls its behavior, and further indicates
conditions of the system. m.sub.1 is the reciprocal hammer mass
that provides the forcing function by delivering the percussion
blow of F force for time duration (t). m.sub.2 is the anvil-bit
mass with its center of mass indicated at X.sub.o, and its one
degree of freedom indicated by an arrow and + or -X. It should be
noted X is the displacement from X.sub.o and + or - indicate
direction and position. For instance, X may be the displacement of
m.sub.2 in the minus direction but it may have a positive or
negative velocity. m.sub.3 is the mass of the drill string that
exerts force W on the anvil thrust ring spring. O is the formation
as a whole. C.sub.2, the damping coefficient of the formation, and
the spring constant, K.sub.2, of the formation are graphical
representations of formation conditions. It is thought that these
variables, at least under some conditions, are velocity dependent
and can be handled with greater accuracy when more knowledge of
rock fracture versus time is available. K.sub.1 and C.sub.1
indicate the spring rate and the damping coefficient of the anvil
thrust ring spring respectively and are design set to compliment
other conditions of the system.
FIG. 18 shows in a comparative graphical way the anvil-bit motions
for the conditions of M.sub.1 and M.sub.2 mode operation. The
curve, C.sub.c, correlates with the force-time graph of FIG. 13 but
curve C.sub.u, which is an example of mode M.sub.2 operation, has
no relation to FIG. 13. This is because the forces of mode M.sub.2
are not simply additive but additive and magnified. It should be
pointed out, however, that the reinforced forces generated by mode
M.sub.2 operation could have been indicated as an increased F force
in FIG. 13. In steady state mode M.sub.2 operation, the average
energy input is equal to the average energy output and the total
energy of the system is constant for constant conditions, but the
system has stored energy in a "potential energy well" that allows
force magnification. This energy is kinetic as the velocity of the
anvil-bit or potential as force storage in the elastic restoring
forces and these energies are constantly changing and interchanging
during the reciprocating anvil-bit motions. These motions are
described by the center of mass of the anvil-bit relative to time
and conditions. X is the displacement of the center of mass of
m.sub.2 from X.sub.o, X.sub.m being the maximum displacement. Time
progresses from the left at F(t).sub.o where time begins at the
percussive blow and progresses to the right through the completion
of one tool cycle of 360.degree. for each mode operation motion
curve. H indicates relative formation hardness and H.sub.2
indicates the hardness at which critical damping occurs while
H.sub.1 is a harder formation and H.sub.3 is a softer formation.
Curve C.sub.c, a dashed line, describes the motion of the anvil-bit
when the system is critically damped and forms a limit or dividing
line between mode M.sub.1 and M.sub.2 operation. Curve C.sub.o, a
dotted line, describes an overdamped system condition and is an
example of mode M.sub.1 operation. The total cycle time, t.sub.3,
is longer and the displacement is greater than that of C.sub.c
indicating lower frequency and greater instantaneous penetration
rate. The center of mass of m.sub.2 only return, or attempts to
return to X.sub.o. In mode M.sub.2 operation, indicated by a solid
line and denoted C.sub.u, the forcing function F(t) displaces the
mass, m.sub.2, a distance, X.sub.m, in formation hardness, H.sub.1,
in the same manner but the instantaneous penetration is less, and a
certain amount of energy is conserved which returns the anvil-bit
across position X.sub.o a distance, -X. Under the influence of the
elastic restoring forces the motion is reversed and m.sub.2 is in a
minus position but begins positive motion. During the positive
motion another percussive hammer impact, F(t) occurs at time
t.sub.1, cycle end, and reinforces the motion of m.sub.2. The force
associated with this motion is substantially increased as compared
with the force associated with that of mode M.sub.1 operation
because of the force reinforcement and resulting magnification. The
force magnification factor described herebefore is shown
graphically in FIG. 19. Curve C.sub.o is descriptive of operational
mode M.sub.1 in that it is characterized by a damping factor of one
or more and C.sub.u is descriptive of operational mode M.sub.2 in
that it is characterized by a damping factor of less than one and
the percussive energy reinforces system dynamic stored energy. It
should be noted that this specification does not cover underdamped
non-resonant system frequency as this condition should be avoided.
In FIG. 18 this condition would be shown as a percussive blow
occuring during negative travel of m.sub.2 on the solid line
curve.
FIG. 19 is a graph in which the magnification factor has been
plotted against the forced to natural frequency ratio for various
values of the damping factor. It may be used to determine the
system mode M.sub.2 steady-state vibration amplitude of the
anvil-bit. When damping is present and provided that it is less
than critical the amplitude passes through a maximum as the forced
frequency, w, is varied. Also it should be noted that as the
damping factor, C/C.sub.c, approaches unity that the resonant
effects are spread over a wider range. If the system is critically
damped as in mode M.sub.1 operation as when C/C.sub.c = 1, the
mass, m.sub.2, does not effect the system and amplitude tends to
decrease with increased forced frequency and would indicate that a
given tool designed to operate in both modes but operating in mode
m, would be operated more advantageously at half or less than
resonant frequency and at or near critical damping.
A preferred configuration of anvil thrust ring is shown in FIG. 11.
This annular single ring spring element is generally rectangular in
section and the surfaces are defined as a portion of a cone. It
could be considered a modified Belleville dish-shaped ring spring
of a high volumetric efficiency. This shape spring has a relatively
large load capacity, small deflection and small height. Under axial
load the unit deflects or distorts in resistance to the load
applied and, thus, absorbs and stores the load energy. Distortion
stresses are a combination of torsion, shear, tension, and
compression and are within the proportional stress-strain elastic
limits of the material used. A single spring element is shown in
FIG. 11, but any number of units may be stacked either in series
for increased deflection and minimum load or with surfaces parallel
for increased load and minimum deflection or in parallel-series for
increased load and increased deflection. Also, it is anticipated
that instead of frustro-conical upper and lower surfaces as
indicated in FIG. 11, these surfaces may be spherically curved and
that this shape tends to add a stiffening effect and may be used to
vary the spring rate. Another type of anvil thrust ring is shown in
FIG. 14. These thrust rings would have a frustro-conical or
spherical coacting interface that when loaded axially cause radial
forces that load the inner ring in compression and the outer ring
in tension causing distortion of said rings and, therefore, energy
storage that can be recovered in the axial direction when the rings
return to their reduced stress conditions as when the anvil is
extended. Yet another configuration of anvil thrust ring spring is
shown in FIG. 9. This unit is made of ring-shaped flat washer metal
plates spaced at intervals with an elastomer molded to and between
the plates. This type anvil thrust ring possibly utilizing urethane
or rubber would have application in the smaller portable mining
blast hole drilling tools where the loads encountered would be
lesser.
Although three preferred types of anvil thrust ring spring
configurations are shown in the Drawings and indicated in FIGS. 2,
22, and 11 by 5, in FIG. 9 by 5B, and in FIG. 14 by 5A, it is
anticipated that other spring designs such as coiled wire can be
used to implement the baic idea involved. The above considers
configuration or shape, materials, heat treatment, number of units
stacked, method of stacking, total loads, rate of load and
deflection, etc. There would be many different requirements of the
anvil thrust rings in the numerous applications and, therefore,
many designs, but all essentially fulfilling the requirements as
before stated.
The basic idea involved in the use of the anvil thrust rings, force
magnification, inphase energy usage from two different sources, and
the superposition of percussive blow force on drill collar weight
applications has been explained throughout this specification, but
the idea is not limited to use with dead weight forces caused by
gravity. Most modern smaller drilling rigs as used for seismograph
shot hole drilling, mining blast hole drilling, core drilling and
quarry boring are equipped with automatic hydraulic pull down units
on the rig that can apply force on the bit should dead weight not
be used, and the same principles are applicable as described
herebefore. Also, the basic idea is not limited to vertical
application, but may be used in any direction as long as the drill
stem is forced in the direction of drilling.
The percussion tool is normally installed in the rotary drill
string above the bit denoted by C and below the drill collars
denoted by A. It is driven by and operable from the drilling fluid
normally circulated in the system for cleaning formation cuttings
from the borehole and for subterranean formation and pressure
control. The drilling fluid under pressure is forced down through
the inside of the drill string, through the percussion tool for
operation of same and exhausts through the bit and flows up the
borehole, outside the drill string to the surface.
As installed in the drill string, lower extendable joint is closed
or retracted due to formation resistance to the bit and drill
collar load applied on the tool housing, said load being
transferred through the housing, anvil thrust rings, anvil, and bit
to the formation. Hammer is in its lower position resting on the
anvil and the valve is positioned to allow fluid flow below the
hammer.
Fluid is introduced to the liquid driven percussion tool, FIGS. 1
and 2, through the central bore in the top sub of the housing,
flows into the upper tubular end of the valve, through its radially
divergent side ports, past the open annular valve seat and valve
face 47, down between outside of the lower tubular valve end and
hammer bore to the substantially closed chamber below hammer,
formed in part by the lower surface of the hammer. Pressure fluid
acts on the downward facing exposed hammer surface causing it to
lift and accelerate the hammer relatively away from anvil. At the
same time, second valve seat and valve face 45 is open and allows
exhaust of fluid and pressure dissipation from the substantially
closed chamber above hammer, formed in part by upwardly facing
surface of hammer. This exhaust fluid travels from said chamber
above hammer through concentrically spaced passages in hammer,
through openly spaced valve and seat surface, through radially
divergent side ports below pressure restriction orifice into lower
tubular end of valve, into bore of anvil and on through the
system.
Hammer and valve move upwardly away from anvil in unison until
hammer strikes upper inside surface of housing at which time hammer
rebounds downwardly and valve, due to its inertia, continues
upwardly until it reseats in its alternate position relative to the
hammer. Passages formed by valve faces and seats that were open are
now closed and alternate passages also formed between annular valve
faces 46 and 48 and their respective coacting seats are open. This
opening of one set of valve surfaces and the closing of the
alternate set of valve surfaces redirect the supply and exhaust
fluid to and from substantially closed chambers on opposite ends of
the hammer. This is to say that pressure fluid is now redirected to
top chamber to act on top surface of hammer and drive it downwardly
relatively toward anvil while exhausting chamber below hammer,
allowing pressure and fluid dissipation therein. The same chambers
and communicating passages are used for fluid redirection but
different valve face and seat surfaces allow passage of fluid or
stoppage of fluid flow.
The hammer and valve again move in unison downwardly. At the
termination of hammer down stroke, hammer strikes anvil a
percussive blow, said blow coacting with stored force in anvil
thrust rings to instantaneously accelerate the anvil-bit, creating
a high impulse force spike that is transferred through the bit to
the formation for crushing same. At percussion and sudden
deacceleration or rebound of hammer, valve, due to its inertia,
shifts to original lowermost downward position relative to hammer
and hammer is again accelerated upwardly as fluid again flows as
orginally directed below the hammer.
An operation cycle of the liquid tool mass motor is now complete
and the tool continues sustained high frequency operation
responding to pressure fluid supplied, formation hardness, etc.,
while being rotated conventionally.
Valve shifts from one position to its alternate position are
unusually fast because of the inertia tending to maintain valve
speed at terminal hammer velocity and also because of conversed
energy in the hammer, causing it to rebound in the direction
opposite of valve inertial travel. Hammer and valve each move a
portion of the short shift travel distance. This quick shift tends
to minimize valve erosion. It is also very important because total
hammer stroke is short and, therefore, relatively less total hammer
stroke is used for valve shift. This means that a greater
percentage of total stroke is used for hammer acceleration and
corresponding increase in cycle frequency.
Also, the fact that the hammer is accelerated to collision before
any valve shift occurs is of major importance in this liquid driven
motor. It insures maximum terminal velocity of hammer at
percussion, since hammer inertia in heavy fluid, as drilling fluid
often is, can be reduced considerably in very short distances. This
not only adds to cycle frequency, but insures maximum production
and transfer of percussive energy. Close dimensional control of
hammer, hammer stroke, housing length and anvil end location is
also eliminated since end of hammer stroke determines valve shift
and hammer reversal.
The gas driven motor valving structure effectively causes hammer
oscillation and anvil-bit excitation just as the liquid valving
structure does but the valving action functions somewhat
differently because of the gaseous state of the fluid. Gas is
introduced to the tool through the top central opening, flows into
the upper tubular end of valve 75, through its side ports 80,
counterbore 83, passages 86 and 42, and into chamber 39 below the
hammer, tending to raise it from its lower position while lower
pressure gas is exhausted from chamber 36 above the hammer which
flows through passages 87 and bores 84 of valve cage, through side
ports 81 and lower tubular end of valve 75 into the bore of the
anvil and on through the system. It is apparent now that a
differential pressure exist across the hammer and that it will move
relatively up away from the anvil and gain velocity and momentum.
The ports in valve 75 and the counterbores 83 and 84 in the valve
cage are spaced so that after a certain valve cage travel all of
the ports are blocked to fluid communication momentarily but
because of hammer momentum and gas elasticity, the hammer continues
to travel upwardly until valve ports have communication with the
opposite chambers at the ends of the hammer. High pressure fluid
now will flow from valve port 80, into bores 84, through passages
87, into chamber 36 above hammer while fluid from chamber 39
exhausts through passage 42 into valve side port 81 exposed below
valve cage and on through the system thus creating a differential
pressure force that will drive the hammer downwardly relatively
toward the anvil. Of course, because valving occurred on the hammer
upstroke the hammer will travel upward a comparatively long
distance before it is reversed by the high pressure fluid acting to
accelerate it downward and bores 84 are rather long to provide full
long stroke driving power for hammer acceleration and percussion
with the anvil. Before percussion occurs fluid pressures are
switched back as originally directed for another hammer upstroke.
Valving is such that it switches intermediate the hammer stroke
ends thus always creating a bias on the hammer and oscillation of
same. This valving may be made simpler because gas is usually
cleaner than liquids and is very tolerant to volume changes and is
very resilient.
During steady-state drilling system operation the drill bit moves
forward through the formation at an average rate that is equal to
the rate of travel of the drill string that follows. The drill bit
although having an average rate of progression is oscillating
axially since the anvil-bit is periodically being driven forward by
the percussive motor and is alternately returned by formation
resistance to the bit. The center of mass of the anvil-bit then is
oscillating relative to its average center of mass location and to
the drill string. Also the anvil thrust rings are storing energy
when the anvil-bit is moving closer to the drill string and are
releasing said stored energy when the anvil bit is moving away from
the drill string and are therefore adding energy to the forward bit
thrusts and absorbing energy from the drill string as the bit slows
its forward movement. The preload of the system which is the
gravitational effect of the drill string or other equivalent force
is continually pumping its energy into the oscillating system and
is used in phase with the forward thrusts of the anvil bit to
increase the amplitude of oscillation. Now depending on formation
conditions which are the primary damping effects and to a large
extent the elastic restoring force of the system, various
conditions of operation are possible. If the system is overdamped
or critically damped, superposition forces can be applied to the
anvil-bit. This is to say that the percussive blow impulsive force
and the anvil thrust ring spring forces are additive resulting in
increased magnitude of forward thrusts of the bit. If the system is
underdamped, magnified superposition forces can be applied for
increased oscillation amplitude and increased formation
disintegration.
The above describes the basic underlying operational parameters of
the percussion tool. The key to this whole vibrational spring-mass
drilling system lies in the anvil thrust ring elastic element(s)
and their ability to store and return energy, their ability to
change the phase of said energy usage, and their ability to coact
with and reinforce the impulsive forces produced by the fluid
driven inertial motor. The system then has the ability to store
energy over comparatively long periods from two different sources,
coordinate it and applies it in comparatively short period
concentrated high energy bursts for increased effectivity in
drilling resulting in an extraordinary drilling device.
These motors are so designed that they will not percuss against
anvil when the lower extendable joint is fully extended. This
permits full pressure bore hole flushing but no percussion tool
operation. When drill string weight is lifted allowing bit and
anvil to extend, hammer and valve are also lowered relative to the
housing. Before full joint extension, top tubular end of valve is
withdrawn from engagement with small central upper bore of housing
allowing fluid to flow therebetween into top chamber through hammer
and valve and on through the tool. As long as valve is out of
engagement with upper bore of housing, piston will not lift, yet
allows full fluid passage.
Although specific embodiments of the invention are illustrated in
the drawings and described herein, it will be understood that the
invention is not limited to the embodiments disclosed, but is
capable of rearrangement, modification, and substitution of parts
and elements without departing from the spirit of the
invention.
* * * * *