[0001] This invention relates to flow pulsing apparatus and a method for use in down-hole
drilling equipment, and in particular to improved apparatus and methods of this type
to be utilized in a drill string above a drill bit with a view to securing improvements
in the drilling process.
BACKGROUND OF THE INVENTION
[0002] In the drilling of deep wells such as oil and gas wells, it is common practise to
drill utilizing the rotary drilling method. A suitably constructed derrick suspends
the block and hook arrangement, together with a swivel, drill pipe, drill collars,
other suitable drilling tools, for example reamers, shock tools, etc. with a drill
bit being located at the extreme bottom end of this assembly which is commonly called
the drill string.
[0003] The drill string is rotated from the surface by the kelly which is rotated by a rotary
table. During the course of the drilling operation, drilling fluid, often called drilling
mud, is pumped downwardly through the hollow drill string. This drilling mud is pumped
by relatively large capacity mud pumps. At the drill bit this mud cleans the rolling
cones of the drill bit, removes or clears away the rock chips from the cutting surface
and lifts and carries such rock chips upwardly along the well bore to the surface.
[0004] In more recent years, around 1948, the openings in the drill bit allowing escape
of drilling mud were equipped with jets to provide a high velocity fluid flow near
the bit. The result of this was that the penetration rate or effectiveness of the
drilling increased dramatically. As a result of this almost all drill bits presently
used are equipped with jets thereby to take advantage of this increased efficiency.
It is worthwhile to note that between 45-65% of all hydraulic power output from the
mud pump is being used to accelerate the drilling fluid or mud in the drill bit jet
with this high velocity flow energy ultimately being partially converted to pressure
energy with the chips being lifted upwardly from the bottom of the hole and carried
to the surface as previously described.
[0005] As is well known in the art, a rock bit drills by forming successive small craters
in the rock face as it is contacted by the individual bit teeth. Once the bit tooth
has formed a crater, the next problem is the removal of the chips from the crater.
As is well known in the art, depending upon the type of formation being drilled, and
the shape of the crater thus produced, certain crater types require much more assistance
from the drilling fluid to effect proper chip removal than do other types of craters.
[0006] The effect of drill bit weight on penetration rate is also well known. If adequate
cleaning of the rock chips from the rock face is effected, doubling of the bit weight
will double the penetration rate, i.e. the penetration rate will be directly proportional
to the bit weight. However, if inadequate cleaning takes place, further increases
in bit weight will not cause corresponding increases in drilling rate owing to the
fact that formation chips which are not cleared away are being reground thus wasting
energy. If this situation occurs, one solution is to increase the pressure of the
drilling fluid thereby hopefully to clear away the formation chips in which event
a further increase in bit weight will cause a corresponding increase in drilling rate.
Again, at this increased drilling rate, a situation can again be reached wherein inadequate
cleaning is taking place at the rock face and further increases in bit weight will
not significantly affect the drilling rate and, again, the only solution here is to
again increase the drilling fluid pumping pressure thereby hopefully to properly clear
the formation chips from the rock face to avoid regrinding of same. Those skilled
in the art will appreciate that bit weight and drilling fluid pressure must be increased
in conjunction with one another. An increase in drilling fluid pressure will not,
in itself, usually effect any change in drilling rate in harder formations; fluid
pressure and drill bit weight must be varied in conjunction with one another to achieve
the most efficient result. For a further discussion of the effect of rotary drilling
hydraulics on penetration rate, reference may be had to standard texts on the subject.
[0007] It should also be noted that in softer formations, the bit weight that can be used
effectively is limited by the amount of fluid cleaning available below the bit. In
very soft formations the hydraulic action of the drilling fluid may do a significant
amount of the removal work.
[0008] In an effort to increase the drilling rate, the prior art has provided vibrating
devices known as mud hammers which cause a striker hammer to repeatedly apply sharp
blows to an anvil, which sharp blows are transmitted through the drill bit to the
teeth of the rolling cones. This has been found to increase the drilling rate significantly;
the disadvantage however is that both the bit life and mud hammer life are significantly
reduced. In a deep well, it is well known that it takes a considerable length of time
to remove and replace a worn out bit and/or mud hammer and hence in using this type
of conventional mud hammer equipment the increased drilling rate made possible is
offset to a significant degree by the reduction in bit and mud hammer life.
[0009] The prior art has also provided various devices for effecting pulsations in the flow
of drilling fluid to enhance the hydraulic action of the drilling fluid and to induce
vibrations in the drill string by virtue of water hammer effect.
[0010] My above-noted copending U.S. Patent Applications Serial Nos. 008963 and 626,121
(disclosures of which are incorporated herein by reference thereto) disclose improved
devices for increasing drilling rate by periodically interrupting the flow to produce
pressure pulses therein and a water hammer effect which acts on the drill string to
increase the penetration rate of the bit. The flow pulsing apparatus described includes
a rotor having blades which is adapted to rotate in response to the flow of drilling
fluid through the tool housing. A rotary valve forms part of the rotor and alternately
restricts and opens the fluid flow passages thereby to create cyclical pressure variations.
The flow passages comprise radially arranged port means in a valve section of the
housing with the rotary valve means being arranged to rotate in close co-operating
relationship to the port means to alternately open and close the radial ports during
rotation.
[0011] Because of the fact that the drilling fluid typically contains a substantial portion
of gritty material of varying size as well as other forms of debris such as sawdust
and wood chips, and since it is not practical to attempt to screen or filter all of
this material out of the drilling fluid, all of the above-described rotary valve arrangements
are somewhat prone to jamming due to debris binding in the valve surfaces. Accordingly,
there is a requirement that a degree of clearance be maintained between the valve
surfaces, and in my above-noted copending applications SN 8963 and 626121 various
improvements have been incorporated thereby to allow the radial clearances between
the valving surfaces to be kept as small as possible while at the same time reducing
the incidence of jamming. It should be kept in mind, of course, that in order to achieve
the maximum water hammer effect, the clearances should be kept as small as possible
thereby to achieve the maximum possible conversion of the flow energy of the drilling
fluid into a water hammer effect. The structures described in my copending U.S. applications
SN 8963 and 626121 require a minimum radial clearance in order to avoid binding and
jamming. Hence, it can readily be seen that the total "leakage" area when the valve
is "closed" will be equal to the clearance dimension multiplied by the total distance
around the valve ports. Since there is a need to keep the total leakage area relatively
small, it follows that the total distance around the valve ports must be kept reasonably
small as well, resulting in much smaller than optimum port holes which in turn restrict
the flow unduly even when the valve is fully open thus creating a substantial pressure
drop across the open valve. This restriction of the flow through the fully open valve
reduces the overall operating efficiency of the system thus tending to restrict its
use for large flow volume situations, i.e. large tools using 400-1100 gallons/minute,
for reasons which will be readily apparent to those skilled in the art.
[0012] My above-noted copending application Serial No. 046,621 describes improved flow pulsing
apparatus adapted to be connected in a drill string above a drill bit and includes
a housing providing a passage for a flow of the drilling fluid toward the bit. A turbine
is located in the housing and it is rotated during use about an axis by the flow of
drilling fluid. A novel valve arrangement operated by the turbine means periodically
restricts the flow through the passage to create pulsations in the flow and a cyclical
water hammer effect to vibrate the housing and the drill bit during use. This valve
means is reciprocated in response to the rotation of the turbine means to effect the
periodic restriction of the flow as opposed to being rotated as in the other arrangements
described above. A cam means is provided for effecting the reciprocation of the valve
means in response to rotation of the turbine means. The cam means preferably comprises
an annular cam surrounding the axis of rotation of the turbine with cam follower means
engaging the annular cam with relative rotation occurring between the follower means
and the cam on rotation of the turbine to effect the reciprocation of the valve. The
valve means includes a valve member which is mounted for reciprocation along the axis
of rotation of the turbine. The axis of rotation, when the flow pulsing apparatus
is located in the drill string, extends longitudinally of the drill string in a generally
vertical orientation.
[0013] By utilizing the reciprocating valve structure described in the above-noted U.S.
application 046,621 a substantial restriction of the flow area is theoretically possible
thus enabling substantial conversion of flow energy to dynamic pressure energy and
achieving a large pressure pulse or water hammer effect. At the same time this novel
valving arrangement is capable of providing a large fluid flow area when the valve
is open thus reducing head losses in the valve full open position and thus in turn
allowing increased throughput of drilling fluid to provide good efficiency. However,
it has been noted that there is a tendency for the turbine in the above arrangement
to stall if the closure or restriction is made very small to achieve the highest water
hammer effect. Stalling is due to the fact that the turbine requires at least some
flow to produce rotation; this means that full closure cannot be achieved in practice
thus limiting the maximum water hammer effect (WHE) achievable.
SUMMARY OF THE INVENTION
[0014] In accordance with the present invention there is provided an improved flow pulsing
method and an apparatus incorporating a movable valve member for producing an enhanced
water hammer effect. This apparatus eliminates the need for the turbine described
in the applications noted above and instead is constructed to set a valve member forming
part of a mass-spring system into oscillation in response to the dynamic forces/vibrations
arising during a drilling operation and/or by the direct action of the drilling fluid
on the mass-spring system thereby to effect intermittent pulsations in the flow thus
achieving the desired water hammer effect. Since this novel method and apparatus do
not employ a turbine, there is no need to maintain a minimum flow through the flow
pulsing apparatus; hence the valve member can close completely during each cycle of
oscillatory motion. This gives rise to a substantially enhanced water hammer effect
(WHE) as compared with the (WHE) achieved by certain prior art arrangements and the
arrangements described in the above-noted patent applications.
[0015] In one form of the invention, the valve member is mounted via suitable guide means
for reciprocation in the axial direction, i.e. lengthwise of the drill string. A spring
is connected to the valve member with the spring and the mass of the valve member
preferably being chosen such that the mass spring system has a resonant frequency
within the range of frequencies of axial vibration likely to be encountered by the
drill string. As described more fully hereafter, the major source of vibration or
displacement is the drill bit itself.
[0016] In another and more preferred form of the invention, a special spring/mass system
is associated with the valve member and the valve member is related to a valve seat
so that it moves against the flow direction to the closing position. The arrangement
is such that pulsation can occur in response to the action of the drilling fluid on
the valve member without the need for drill string oscillation. The shape of the pulses
and pulse frequency can be preselected to some degree by altering the mass or spring
constant etc. of the spring-mass system. When the frequency of the spring-mass system
is chosen to be close to the natural frequency of the rest of the drill string (or
the bottom part of the string when isolated by a shock tool or other telescopic member
from the string above) the spring-mass system can oscillate in resonance with the
drill string (or part of it) with the result being that enormous amounts of energy
are transmitted to the bit. The arrangement is also resistant to clogging due to debris
and since the valve opens in the flow direction, if the spring breaks the valve merely
stays open continually thus permitting drilling to continue (at a slower rate) and
deferring a costly trip out of the hole.
[0017] Further features of the invention and the advantages associated with same will be
apparent to those skilled in the art from the following description of preferred embodiments
of the invention when read in conjunction with the accompanying drawings.
BRIEF DESCRIPTION OF THE VIEWS OF DRAWINGS
[0018]
Figure 1 is a graph illustrating the relationship between drilling rate and bit weight
and illustrating the effect that increased cleaning has on drilling rate;
Figure 2 is a longitudinal section at the bottom of a well bore illustrating apparatus
according to the invention connected in the drill string immediately above the drill
bit;
Figure 2A is a modification of the arrangement shown in Fig. 2;
Figure 3 is a diagrammatic view of the bottom end of the well bore illustrating a
jet of drilling fluid being emitted toward the wall and bottom of the bore hole;
Figure 4 is a longitudinal half section of apparatus for producing a pulsating flow
of drilling fluid in accordance with a first embodiment of the invention;
Figure 5 is a cross-section view taken along line 5-5 of Fig. 4;
Figure 6 is a longitudinal half section of a second embodiment of the flow pulsing
apparatus;
Figure 6A is an enlarged view of a portion of Fig. 6;
Figure 7 is a hypothetical pressure - time plot taken above the valve means;
Figures 8 and 9 are pressure - time plots taken above and below the valve means of
the embodiment of Figure 6; and
Figure 10 is a plot of spring force - valve member displacement for the Figure 6 embodiment.
Figure 11 is a longitudinal half section of a third embodiment of the apparatus, similar
to the embodiment of Figure 6 but of somewhat simplified form.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0019] Reference will be had firstly to Fig. 1. As noted previously the effect of bit weight
on penetration rate is well known. With adequate cleaning, penetration rate is directly
proportional to bit weight. There are some limitations depending of course upon the
type of formation being drilled. There is also, in any particular situation, a maximum
upper limit to the magnitude of the weight which the bit can withstand.
[0020] With reference to Fig. 1, it will be seen that drilling rate is generally proportional
to bit weight up to point A where drilling rate drops off rapidly owing to inadequate
cleaning which means that formation chips are being reground. From point A, increased
cleaning resulted in a proportional increase in drilling rate up to point B where,
again, inadequate cleaning was in evidence with a consequent fall off in drilling
rate. Again, by increasing the cleaning effect, drilling rate once again became proportional
to bit weight up to point C where again, a fall off in drilling rate is in evidence.
[0021] Fig. 1 thus demonstrates clearly the importance of effective hole bottom cleaning
in obtaining an adequate drilling rate.
[0022] It is noted that Fig. 1 has been described mainly in relation to the drilling of
harder formations. In softer formations, where the hydraulic action of the drilling
fluid does at least part of the work, the relationships shown in Fig. 1 would still
apply, although for somewhat different reasons, as those skilled in the art will appreciate.
[0023] Referring now to Figure 2, there is shown in cross section the lower end portion
of a bore hole within which the lower end of a drill string 10 is disposed, such drill
string including sections of hollow drill pipe connected together in the usual fashion
and adapted to carry drilling fluid downwardly from drill pumps (not shown) located
at the surface. The drill string is driven in rotation by the usual surface mounted
equipment also not shown. Attached to the lower end of the drill collar 12 via the
usual tapered screw thread arrangement is a drilling fluid flow pulsing apparatus
16 in accordance with the invention. To the lower end of the flow pulsing apparatus
is connected a relatively short connecting sub 18 which, in turn, is connected via
the usual screw threads to a drill bit 20 which may be of conventional design having
the usual rolling cone cutters and being equipped with a plurality of cleaning jets
suitably positioned to apply streams of drilling fluid on to those regions where they
have been found to be most effective in removing chips from the bottom of the well
bore. A somewhat modified arrangement is shown in Figure 2A wherein, above the flow-pulsing
apparatus 16, there is provided a drill collar section 17 (to provide extra mass)
and above that, a telescoping section 19 of conventional construction which can isolate
the upper part of the drill string from the bottom section. The usual rolling cone
cutters can be replaced with a percussive bit when the flow pulsing is in a resonant
relationship to the rest of the drill string or in reasonance with the lower end of
the drill string (when the isolating telescopic member 19 (eg. a standard bumper sub
or shock tool)) is interposed above the flow pulsing apparatus 16 as shown in Figure
2A. One of such cleaning jets 22 is diagrammatically illustrated in Fig. 3 (the remainder
of the drill bit not being shown) thereby to illustrate the manner in which the jet
of drilling fluid is directed against the side and bottom portions of the well bore
during a drilling operation. The location and arrangement of the jet openings on the
drill bit 20 need not be described further since they are not, in themselves, a part
of the present invention but may be constructed and arranged in an entirely conventional
manner.
[0024] Referring now to Figs. 4 and 5, the first embodiment of the flow pulsing apparatus
16 is shown in detail. Apparatus 16 includes an external tubular casing 26, the wall
of which is sufficiently thick as to withstand the torsional and axial forces applied
thereto during the course of the drilling operation. Casing 26 is in two sections
which are connected together via tapered screw threaded portion 28, with the upper
end of the casing having a tapered internally threaded portion (not shown) adapted
for connection to a lower end portion of the drill string. The casing 26 also includes
a tapered internally threaded lower section (not shown) which may be connected to
the drill bit 20.
[0025] The casing 26 has a removable cartridge 32 located therein, cartridge 32 containing
the valve means to be hereafter described.
[0026] The cartridge 32 includes an outer cylindrical shell 34. An elongated valve guide
36 is supported co-axially in shell 34 by means of radial fins 38 interconnected between
the interior of shell 34 and the guide 36.
[0027] The upstream end 40 of guide 36 is of relatively small diameter; the downstream end
is of larger diameter and comprises a sleeve 42 of very hard material, e.g. tungsten
carbide, sleeve 42 being connected to intermediate section 46 which, in turn, is fixed
to upstream end 40. The upstream end 40 is provided with a smooth conical nose 48
which directs the flow of drilling fluid around the guide 36.
[0028] An axially movable valve member 50 is located in the valve guide 36 for axial movement
therein and it includes a large head end 52, a small stem portion 54, and an intermediate
section 56. A coil compression spring 60 surrounds the stem 54 and its one end bears
against a ring 62 affixed to the end of stem 54 by pin 64, while the other end of
spring 60 bears against an annular stop 66 fixed to guide upstream end portion 40.
An inner annular bearing portion 70 extends between stop 66 and the interior of sleeve
42 and the downstream end of bearing 70 has a shoulder 72 defining the upstream limit
of travel of valve member 50.
[0029] Valve member 50 has drilled apertures 74, 76 therein allowing the drilling fluid
to have access to both sides of the valve member. The hydraulic forces acting on the
valve member thus act to balance and to cancel one another out.
[0030] The downstream end of shell 34 has an annular valve ring holder 80 seated therein
and held in place by abutment against a step 82 in the casing 26. Holder 80 defines
conical upstream and downstream faces 84, 86 and has an annular step therein which
seats an annular valve ring 88 (and held in place by conical wear ring 89), the valve
ring 88 being co-axially arranged with respect to the valve member 50. Hence, as valve
member 50 moves axially back and forth within the valve guide 36, the head end 52
moves toward and away from the valve ring 88, thus opening and closing the annular
flow passage defined between the head of the valve and the valve ring 88. On the subject
of wear it might be noted that the valve ring 88 is preferably of tungsten carbide
while the valve member 50 is suitably hard-surfaced to avoid excess wear thereof.
(The valve sleeve 42, as previously noted, is preferably of tungsten carbide.) All
other components subject to the abrasive drilling fluid are likewise hard-surfaced
to reduce wear.
[0031] The coil compression spring 60 and the mass of the valve member 50 are chosen so
that the mass-spring system defined by the two of them has a resonant frequency within
the range of the exciting or forcing frequencies arising from the action of the drill
bit on the bore hole bottom. In this regard, reference is had to U.S. Patent 3,307,641
of March 7, 1967 to J.H. Wiggin Jr. which describes in some detail the vertical displacements
of the drill string and frequencies thereof arising from the action of the rolling
cone cutters on the hole bottom. Conventional rolling cone cutters can be used although
special designs can be provided to enhance the displacement as described in the Wiggin
Jr. patent. By rotating the drill string at a selected speed, the vertical displacements
can be of a frequency corresponding to the natural vibrational frequency of the drill
string. Hence the mass-spring system defined by valve member 50 and spring 60 can
be forced to oscillate at that same frequency thus generating pressure pulses (due
to the water hammer effect) in step with the natural vibrational frequency of the
drill string and reinforcing the same. The response of the above mass-spring system
will of course be enhanced if its natural frequency equals the forcing frequency,
i.e. the frequency of the vertical longitudinal displacements of the drill string.
Since the amplitude of the oscillations of the valve member 50 depends to some extent
on the relationship between the natural frequency and the forcing frequency, the head
52 of the valve member 50 is of slightly smaller diameter than the aperture in the
valve ring 88 so that it can enter into such aperture as the amplitude of the oscillations
increase. This permits the valve member to have the desired excursion while eliminating
hammering of the valve member on a seat, which hammering could disrupt the free oscillatory
motion of the valve member and cause wear of the valve members.
[0032] In the embodiment of the invention shown in Figure 6 and 6A (which is a more preferred
form of the invention), the flow pulsing apparatus includes an external casing 100
as before, in two sections, connected by screw threaded portion 102, the upper end
having internally tapered threaded portion 104 adapted for connection to the lower
end of a drill string (not shown) while the lower internally threaded portion 106
may be connected to a drill bit (not shown) via a connecting sub.
[0033] The casing 100 has a removable cartridge 110 therein which contains the valve means
to be hereafter described. Cartridge 110 includes an outer cylindrical shell 112 in
which an elongated valve guide assembly 114 is co-axially supported by means of several
radial fins 116 interconnected between the interior of shell 112 and guide assembly
114. An axially movable valve member 118 is slidably mounted on the upstream end of
guide assembly 114 for movement toward and away from valve seat assembly 120 located
in the upstream end of cartridge 110 and held in place by virtue of mating screw threads
121 on both the seat assembly 120 and the cartridge 110. An annular flow passage is
defined between the valve member 118, guide assembly 114, and the interior of the
shell.
[0034] Valve seat assembly 120 includes an annular ring holder 124 which butts up against
the step 122. Valve ring 126 seats in the ring holder and defines a central throat
128 and opposed, conical, upstream and downstream faces 130, 132, the downstream face
132 defining a valve seat. Valve ring 126 is of very hard material, preferably of
tungsten carbide, and is held in place by an annular step on the holder 124 and by
an annular valve ring holder 134.
[0035] The upstream end of valve member 118 includes a tapered section leading to a reduced
diameter portion 136 which, in turn, leads into a frustro-conical valve face 138 which
cooperates with face 132 of valve ring 126 to prevent flow through the valve when
the valve member 118 is at the upper limit of its travel. The upstream end of valve
member 118 also includes an axially disposed valve tip 140 which extends into the
throat 128 of the valve ring when the valve member 118 approaches the closed position.
The valve tip is of very hard material, e.g., tungsten carbide, and has a rounded
conical nose to meet and divert the flow around the valve member 118 when the latter
is at least partly open.
[0036] Valve tip 140 acts to prevent heavy impact or hammering between the above-noted value
faces 132 and 138, which impacts would shorten valve life span. Tip 140 meets the
incoming flow and by virtue of its close but non-binding fit in the throat of the
valve ring 126, the water hammer effect (WHE) is achieved and equilibrium (to be described
later) is reached in the absence of heavy hammering contact between those faces 132,
130 thus increasing valve life. This is a significant factor especially when it is
considered that the frequency of oscillation of the valve body 118 is likely to be
somewhat greater than 20 Hertz.
[0037] Returning now to the guide assembly 114, the latter includes a tubular upstream barrel
portion 142 which communicates with a downstream elongated tubular spring holder 144.
A bearing sleeve 146 which is preferably of low friction plastics material, e.g.,
nylon, slidably surrounds the barrel and is fixed to the interior bore 148 of valve
member by suitable lock rings, there being a rubber wiper ring 150 at each end of
this sleeve, which rings bear on the outer (polished) surface of barrel 142 to help
clean away grit, etc., thus allowing the valve member 118 to reciprocate freely in
the axial direction along the barrel.
[0038] The spring holder 144 has a spring stop ring 152 at the downstream end thereof against
which an elongated first coil spring 154 bears. This spring 154 extends all the way
to the upstream end of the barrel 142 and makes contact with an axially movable annular
spring support 156, the latter having a tubular portion which fits freely into the
interior of the barrel 142 and against which the upstream end of spring 154 bears;
(the first coil spring has a relatively low spring constant). Spring support 156 is
axially movable relative to both the barrel 142 and the valve member 118 and it has
an annular flange 158 at its upstream end.
[0039] A second relatively short spring 160 (of relatively high spring constant) bears at
its one end against the flange 158 of spring support 142 and at its other end against
a ring 162 which is fixed to the upper interior end of the bore in the valve member
118. As the valve member 118 moves downwardly to open the valve, the first spring
154 (of lower spring constant) is gradually compressed as the spring support 156 moves
along the barrel until the flange 158 contacts the upper terminal end 159 of the barrel.
Further downward movement of the valve member 118 causes compression of the second
spring 160 (of high spring constant). The several parts are dimensioned such that
the total stroke length of the valve member 118 is relatively short (e.g., less than
one inch) in a typical case. In operation, to be described later, most of this movement
results in compression of the first spring 154 while only a small amount (if at all)
of this motion acts to compress the second spring 160.
[0040] Some typical dimensions will be given to help illustrate the operation of the invention,
it being realized that these are not limiting on the scope of the invention but are
given by way of example only:
A. Weight (mass) of valve member (118) - 25 lbs. (11.3 kg)
B. Length of first spring (154) in the installed extended state - 18 ins. (45.7
cm) approx.
C. Length of second spring (160) in the installed extended state - 2 ins. (5.1
cm) approx.
D. Spring constant of first spring (154) - 20 lbs./in (35 Nt/cm) approx.
E. Spring constant of second spring (160) - 1500 lbs./in (2635 Nt/cm) approx.
F. Axial preloading of springs (154 & 160) in the installed extended condition -
80-85 lbs (356-378 Nt) approx.
G. Diameter of throat (128) defined by valve ring (126)- 1 in (2.54 cm)
H. Length of stroke of valve member - 1 in (2.54 cm) max. (approx.)
(i) Amount of compression of spring (154) - (3/4) in (1.92 cm) max. (varies)
(ii) Amount of compression of spring (160) - (1/4) in (.64 cm) max. (varies)
I. Pulse frequency at equilibrium - (25) Hertz approx.
[0041] In the operation of the apparatus of Figure 6, the flow of drilling fluid is accelerated
as it moves downwardly through the throat 128 defined by the valve ring 126. At the
same time, the pressure in this area is reduced due to the Bernoulli effect. The serially
arranged springs 154 and 160 urge valve member 118 and its valve face 138 and tip
140 against the direction of the flow, the preloading in these springs being slightly
greater than the dynamic pressure arising from the flow. Hence, the valve member 118
tends to move in the closing direction until the flow is restricted and the pressure
on the upstream side of the valve increases, such increased pressure acting on the
valve member 118 to cause it to open. At this point, it is noted that the energy (work
done on the valve by the flow as it opens) is stored in the mass/spring system during
opening and is used to overcome the pressure rise above the valve during the closing
of the valve. When the valve closes or severely restricts the flow the (WHE) is achieved.
The increased pressure above the valve acts on the valve spring-mass system and all
the energy (work) required to drive the mass-spring system downwards is stored in
the mass-spring system for use in the next valve closing cycle. The large mass of
the valve member acts as a "flywheel" to store energy during opening of the valve
and this energy is in turn used during closing of the valve.
[0042] The valve closing force is thus proportional to the amount of energy (momentum) that
can be stored in the spring-mass system during opening of the valve and the original
preload on the springs. The result after start-up is that on each successive closing
cycle, the closing force is slightly greater than before thus resulting in a progressively
greater restriction of the valve opening and thus producing higher pressure pulses
due to the water hammer effect (WHE). This build-up continues until:
(a) equilibrium is reached; and
(b) valve member (face 138) comes in contact with face 132 resulting in maximum flow
restriction and maximum (WHE).
[0043] Tests have confirmed the above statements.
[0044] The reasons for making first spring 154 of low spring constant and second spring
160 of high spring constant will now be described. The terms "high" and "low" are
relative terms. The following discussion will help to clarify what is meant by these
terms and will enable those skilled in the art to select spring constants for the
springs which will accomplish the desired result without undue experimentation for
any given situation.
[0045] If the spring constant of spring 154 were made "high", the movement of the valve
member 118 down from the closed position would be very limited (i.e. the stroke would
be short) and all energy from the valve opening pulse would be absorbed quickly and
the valve member would move quickly back to the valve closed position. The graph of
the resulting pressure pulse (WHE) would be as in Figure 7. The pressure differential
to operate the tool would be relatively high (a thousand p.s.i. (7000 kPa) or more)
and the mean pump pressure (MPP) woud be excessively high thus resulting in excessively
high pumping power requirements.
[0046] On the other hand, when the constant of spring 15A is made low, energy storage in
the mass/spring system during the opening stroke will take place over a much longer
stroke than in the previous case thus resulting in a longer time period that the tool
is fully open. The graph of the resulting pressure pulses (WHE) appears as in Figure
8. It can be seen from this that by using a low spring constant for spring 154 the
pulses are well separated or spaced out. The pressure difference (200 psi 1380 kPa
or so) to operate the tool is low and the mean pump pressure (MPP) is also lower,
thus reducing pumping power requirements and a relatively low frequency pulse rate
(eg. 20-27 Hertz) is provided.
[0047] The combined effects of the two springs 154 and 160 will now be described. In order
to further accelerate the return of the valve member 118 to the closed position once
separation of pulses has been achieved by use of the low spring constant spring 154,
use is made (in the Fig. 6 embodiment) of the high spring constant spring 160. This
spring 16 is effectively activated toward the end of the opening stroke of valve member
118 when the flange 158 on movable spring support 156 engages with the top end 159
of the barrel 142 on which the valve member is mounted. Once this second spring 160
starts compressing during the latter part of the stroke of the valve member, all remaining
energy from the opening impulse is stored over a very short portion of the stroke
and the valve member is returned more quickly up to the closed position. In other
words, the use of the high spring constant spring 160 creates greater acceleration
of the valve member 118 toward the closing position thus resulting in a somewhat higher
pulse frequency while at the same time the separation of the pulses and the advantages
associated therewith, e.g., lower pressure differential and (MPP) as outlined above
in connection with Figure 8 are maintained.
[0048] It is not easy to define with precision the preferred relation between the spring
constants of the two springs 154 and 160. In the example given above, the ratio of
the high to the low spring constant is 1500 lb/in (2625 Nt/cm) : 20 lb/in (35 Nt/cm)
or 75. This ratio can be varied substantially, e.g., from 50 to 90 and possibly as
much as 25 to 100 depending on the precise application. Hence, the expressions "high"
and "low" spring constants are used here to describe the fact that the constant of
one spring can be many times higher, (in most cases several order of magnitudes higher),
than that of the other spring. It is also noted here that the second spring can be
dispensed with altogether and a further embodiment to be described hereafter omits
the second spring.
[0049] In common with the first embodiment of the invention described in connection with
Figures 4 and 5 it is possible to operate the embodiment of Figure 6 in a resonant
mode if the natural frequency of the valve spring-mass system is made to match the
natural frequency of the drill string or the natural frequency of a bottom end of
a drill spring that is isolated from the upper end of the drill string by a telescopic
member, shock sub or the like (Fig. 2A). However, the embodiment of Figure 6 need
not be used with a bit capable of producing significant vertical displacements of
the drill string, e.g., it is capable of pulsating on its own independently of any
oscillation of the drill string. When used in a drill string which is vibrated axially
by the bit, the embodiment of Figure 6 would be self-starting in the sense that it
would begin to pulse the flow independently; however, once the suspended mass of the
drill string (e.g., drill bit, flow pulsing apparatus and male spline of a stock tool,
if present) begin to oscillate, then the mass/spring system defined by the valve member
118 and its springs will begin to oscillate and the whole oscillating assembly can
be made to oscillate in resonance.
[0050] It can hence be seen that the embodiment of Figure 6 is more versatile than the first
embodiment (Figures 4 and 5). It (the Figure 6 version) is also less prone to jamming
or choking as a result of debris in the flow of drilling fluid (mud) since the valve
member closes in a direction opposite to the flow direction and any particles wedging
between the valve faces, etc., on one closing cycle are usually relieved and swept
away on the next opening cycle.
[0051] The embodiments of Fig. 11 is similar to the embodiment of Fig. 6 and includes a
casing 200 as before with internally threaded upstream and downstream portions 204
& 206. A guide and support assembly 214 includes an elongated barrel 242 supported
by sleeve 270, radial fins 216 and barrel holder 244. A massive valve member 218 (including
its upstream nose sections 272, 273) is mounted for reciprocation on the barrel 242
as before via bronze or plastic brushings 246a and intermediate bronze brushing 246b.
[0052] An elongated spring 254 extends within the barrel 242 from downstream spring stop
252 up to an internal sleeve 270 which is fixed to the forward end section 272 of
valve member 218 and it slides within the end of barrel 242 as the valve member reciprocates
under the influence of the forces described previously.
[0053] The valve ring 226 is mounted in an annular recess defined by the two-part ring holder
224a and 224b. A small amount of clearance in the axial direction is provided between
the valve ring 226 and the two-part valve holder 224(a&b). A rubber shock absorbing
ring 278 is provided between the holder portion 224b and a step defined by the upstream
casing portion 201. Hence, during operation, as the valve ring 226 moves downstream
slightly. As the valve member 218 moves upstream and the valve faces 232, 238 begin
to close on each other the valve ring 226 moves upstream against the hydraulic pressure
that builds up above the valve; after this clearance has been taken up, impact forces
between the valve faces 232, 238 are absorbed in part, by the rubber shock absorbing
ring 278.
[0054] The embodiment of Figure 11 requires only a single spring 254 and the spring mass-system
defined by it and the valve member 218 function as described above in connection with
the Figure 6 embodiment except that the frequency of operation is somewhat lower owing
to the absence of the second (high spring constant) spring. The embodiment of Figure
11 may in fact be the preferred embodiment for many applications.
[0055] During operation of the embodiments described above, the pulsating pressurized flow
being applied to the cleaning nozzles or jets of the drill bit provides greater turbulence
and greater chip cleaning effect than was hitherto possible thus increasing the drilling
rate in harder formations. In softer formations where the eroding action of the drill
bit jets has a significant effect, the pulsating, high turbulence action also has
a beneficial effect on drilling rate. By making use of the water hammer effect, these
high peak pressures are attained without the need for applying additional pumping
pressure at the surface thus meaning that standard pumping pressures can be used while
at the same time achieving much higher than normal maximum flow velocities and pressures
at the drill bit nozzles.
[0056] In the embodiments described above, owing to the water hammer effect created as a
result of the pulsating flow of drilling fluid, mechanical vibrating forces will be
applied to the flow pulsing apparatus which will act in the direction of the drill
string axis, which pulsing or vibrating action will be transmitted to the drill bit.
This pulsating mechanical force on the drill bit complements the pulsating flow being
emitted from the drill bit jet nozzles thereby to greatly enhance the effectiveness
of the drilling operation, i.e. to increase the drilling rate.
1. Flow pulsing apparatus adapted to be connected in a drill string above a drill
bit and including a housing providing a passage for a flow of drilling fluid toward
the bit, and valve means for periodically restricting the flow through said passage
to create pulsations in said flow and a cyclical water hammer effect to vibrate the
housing and the drill bit during use, said valve means including a valve member located
in the flow passage and forming a part of a mass-spring system supported and arranged
for oscillation in response to forces arising from the action of the drilling fluid
on the valve member and/or longitudinal vibrations of the drill string occurring during
use thereby to effect said periodic restriction of the flow.
2. Apparatus according to claim 1 including spring means associated with said valve
member and defining therewith said spring-mass system which oscillates in response
to direct action of the drilling fluid on the valve member and/or longitudinal vibrations
arising in the drill string during use.
3. Apparatus according to claim 2 including means guiding and supporting said valve
member for oscillation along an axis, with said axis of oscillation, when said apparatus
is located in a drill string, extending longitudinally of the drill string.
4. Apparatus according to claim 3 wherein said valve member is so arranged that, during
use, it is bathed in drilling fluid so that the resulting hydraulic pressure forces
on said valve member substantially balance and cancel each other out.
5. Apparatus according to claim 3 wherein said valve means includes an annular ring
fixed to said housing and surrounding said axis of oscillation, said valve member
being arranged such that an annular flow passage is defined between itself and said
ring, said valve member, in use, oscillating along the axis of oscillation toward
and away from said annular ring such that the area of the annular flow passage defined
between said ring and valve member varies from a maximum to a minimum.
6. Apparatus according to claim 1, 2, 3 or 4 wherein said spring means is a coil compression
spring.
7. Apparatus according to claim 1, 2, 3 or 4 wherein said throat includes a ring defining
a central flow passage and said portion of the valve member being adapted to closely
approach or enter into the central flow passage in close proximity to the ring to
effect the restriction or interruption of the flow.
8. Apparatus for effecting pulsations in a flow of drilling fluid through a drill
string whereby to create a cyclical water hammer effect in said drill string, and
comprising:
means defining a passage for flow of drilling fluid;
a valve member located in the flow passage and adapted for oscillating motion in a
direction axially of the drill string when in use;
means guiding and supporting said valve member for oscillating motion in the axial
direction;
spring means connected to said valve member and adapted to be extended and retracted
as the valve member oscillates; said valve member and spring together defining a spring-mass
system adapted for oscillation in response to dynamic forces acting thereon during
use;
means defining an axially disposed throat through which, in use, drilling fluid passes
toward a drill bit; and
said valve member including a portion co-operative with said throat to cyclically
restrict or interrupt the flow therethrough as the valve member oscillates without
disrupting the oscillation thereof.
9. Apparatus according to claim 8 wherein said means for guiding and supporting the
valve member includes an elongated chamber, said valve member having an elongated
stem portion arranged for free axial movement in said chamber, said spring means being
a coil spring connected to said stem portion and surrounding the same and located
in the elongated chamber.
10. Apparatus according to claim 9 wherein said portion of the valve member co-operative
with said throat comprises an enlarged head portion, a portion of which is slidably
located within said elongated chamber.
11. Apparatus according to claim 8 wherein said valve means has a passage therein
allowing hydrostatic fluid pressures to equalize on upstream and downstream sides
of said valve means such that the latter is hydraulically neutral.
12. A rotary percussive drill string assembly comprising:
an elongated tubular drill string having a drill bit capable of imparting axial vibratory
displacements to the drill string on rotation of the drill bit against a borehole
bottom, said drill string being capable of conducting a flow of drilling fluid axially
therealong toward said drill bit to clear away cuttings and the like, said drill string
assembly having therein, above said bit, an apparatus as defined in any one of claims
8-11 with said spring-mass system defined by the spring means and valve member having
a resonant frequency within the range corresponding to the frequency of the axial
vibratory motion of the drill string induceable by said drill bit during a drilling
operation thus to effect periodic pulsations in the flow of fluid passing along the
drill string and a resulting periodic water hammer effect creating periodic axial
forces on the drill bit to enhance the drilling rate.
13. A method of drilling a well comprising rotating within a borehole an elongated
tubular drill string having a drill bit which imparts axial vibratory displacement
to the drill string as the drill bit rotates against the borehole bottom, and passing
drilling fluid through said drill string to said drill bit to clear away cuttings,
and providing in said drill string an apparatus as defined in any of claims 1-11 whereby
to effect oscillation of the valve member by virtue of the vibratory displacement
of the drill string thus causing pulsations in the flow of drilling fluid and a resulting
periodic water hammer effect which, in turn, creates periodic forces on the drill
bit to enhance the drilling rate.
14. Flow pulsing apparatus adapted to be positioned in a drill string above a drill
bit and including a housing providing a passage for a flow of drilling fluid toward
the bit, and valve means for periodically restricting the flow through said passage
to create pulsations in said flow and a cyclical water hammer effect to vibrate the
drill string and the drill bit during use, said valve means including a valve member,
means guiding and supporting said valve member for oscillation along an axis, with
said axis of oscillation, when said apparatus is located in a drill string, extending
longitudinally of the drill string, and spring means associated with said valve member
and defining therewith a spring-mass system which oscillates during use to effect
said periodic restriction of the flow, said valve means for periodically restricting
the flow being arranged such that in use, the oscillating valve member moves (a) axially
opposite to the flow direction toward a flow restricting or closed position and (b)
axially in the flow direction toward an open or non-restricting position.
15. Apparatus according to claim 14 including one or more of the following features
in suitable combination:
A. said spring-mass system is arranged such that said valve member is moved toward
the restricting or closed position by the energy stored in the spring-mass system
during the previous opening movement of the valve member, said valve member being
exposed to the flow of drilling fluid during use and responding to the direct action
of the fluid forces thereon during use;
B. said valve means comprises an elongated valve member having an interior bore therein,
and said guiding and supporting means comprising an elongated guide fixed to said
housing and disposed within the bore in the valve member such that the latter is slidable
thereon during its stroke of travel, and said spring means extending, in part, axially
along said guide and acting on said valve member to urge the latter toward a closed
or flow restricting position;
C. said spring means comprises a pair of springs arranged in series, a first one of
said springs being of a relatively low spring constant to provide for a desired natural
rate of frequency of oscillation while the second one of said springs is of a relatively
high spring constant and is arranged to be activated during the latter part of the
opening stroke of the valve member to effect a relatively rapid return of the valve
member to the flow restricting position whereby to provide separation of the pulsations
in the flow;
D. an axially movable spring support located at an upstream end of the guide and interposed
between the first and second springs, said spring support cooperating with said guide
to allow compression of the first spring during a first major portion of the opening
stroke of the valve member and compression only of said second spring during a second
minor portion of said stroke;
E. said spring means comprises first and second springs arranged in series along said
axis of oscillation, said second spring having a spring constant substantially higher
than that of the first spring, and means cooperating with said first and second springs
to (a) allow compression of the first spring during a first portion of the movement
of the valve member in the closing direction and (b) allow compression of the second
spring only during a second portion of the movement of the valve member in the closing
direction;
F. said valve means includes an annular ring fixed to said housing and surrounding
said axis of oscillation, said valve member being arranged such that an annular flow
passage is defined between itself and said ring in the open position of said valve
member, said valve member, in use, oscillating along the axis of oscillation toward
and away from said annular ring such that the area of the annular flow passage defined
between said ring and valve member varies from a maximum to a minimum;
G. said valve member and said valve ring define mating annular valve seats, said valve
ring defining a circular throat portion and said valve member having a tip portion
thereon which enters into the throat before said valve seats contact each other whereby
forces arising from the dynamic pressure of the flow of drilling fluid act on said
tip portion to reduce the speed of movement of the valve member and any impact between
the valve faces.