Note: Descriptions are shown in the official language in which they were submitted.
5S~.
-- 2 --
~CKGROUND OF THE INVENTION
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.
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.
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
3~ fluid or mud in the drill bit jet with this high velocity
Slow energy ultimately being partially converted to
pressure energy with the CllipS being lifted upwardly from
the bottom of the hole and carried to the surface as
previously described.
As is well known in the art, a rock bit drills by
forming successive small craters in the rock face as it is
~2~ S~
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 are,
depending upon the type of formation being drilled, and
the shape of the crater thus produced, certain crater
types require ~uch more assistance from the drilling fluid
to effect proper chip removal than do other types of
craters. For a further discussion of this see "Full Scale
Laboratory Drilliny Tests" by Terra-Tek Inc., performed
under contract Ey-76c-024098 for the U.S. Department of
Energy.
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, iOe. 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 i5 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 th~ 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 ancl drilling fluid pressure must be increased in
conjunction with one another. An increase in drilling
fluid pressure will not, in itself, usually effect any
sss~
-- 4 --
change in drilling rate in harder formations; ~luid
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.
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 avai-lable below
the bit. In very soft formations the hydraulic action of
the drilling fluid may do a significant amount of the
removal work.
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, wllich 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 the bit
life is 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 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 life.
One proposal for cyclically interrupting flow
through a drill stem is disclosed in U.SO Patent No.
2,780,438 issued February 5, 1957. This patent proposes
the use of a rotary value member ac-tuated by a spiral
rotary value actuator. Axially disposed co-operating
passages are provided in the valve structure and thrust
bearings take up axially oriented loads on the rotary
valve ~ember. Disadvant~ges of this proposal include the
fact that the axially oriented passages are prone to
blockage by debris. The high shock forces on the rotary
valve member would tend to rapidly destroy the thrust
~ ~il5SSl
-- 5 --
bearings supporting the rotary valve. The overall
arrangement would be very inefficient in providing
fluctuating forces on the drill bit. The free telescoping
movement of the housing above the rotary valve would
destroy most of the desired water hammer effect and would
appear to eliminate most of the pressure drop below the
bit considering that the apparatus is acting in a closed
system.
Another prior art flow pulsing arrangement is
shown in the Zublin U.S. Patent 2,743,083 issued April 24,
1956. This patent shows several embodiments of an
invention. ln all of these embodiments, however, the
arrangement is such that pressure pulses above the rotor
and consequent pressure drops below the rotor act on
almost the whole projected area of the rotor. High axial
forces on the rotor bearings result thus materially
shortening the bearing life. Furthermore, ~he valving
arrangements provided are prone to jamming due to debris
in the drilling fluid and if sufficient clearance is
provided to alleviate jamming problems the structural
configuration of the valve ~akes it difficult to achi~ve a
meaningful level of pressure build-up.
CA~ 59
My above-noted ~k~ Patent 478~4T~5 discloses improved
forms of flow pulsing apparatus including a rotor having
blades which i adapted to rotate in response to the flow
of drilling fluid through the tool housing. A rotary
valve foxms 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 ~eans 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.
~2~3~S~i~
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 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 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 avoiding jamming
under ordinary circumstances. 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
dynamic pressure energy to produce the optimum water
hammer effect. The structures described in my copending
applications above 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 ~urn 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 for reasons which will be readily
apparent to those skilled in the art.
;5~
Another disadvantage associated with rotary valve
flow pulsating arrangements is that the timing or
~requency of the fluctuation is strictly governed by the
angular velocity of the rotor. Another disadvantage is
that the shape of the pressure pulse curve cannot be
easily varied or changed to better suit conditions.
SUMMARY OF T~IE IMVENTION
The present invention provides improved flow
pulsing apparatus adapted to be connected in a drill
string above a drill bit and incluaes a housing providing
a passage for a flow of the drilling fluid toward the
bit. A turbine means 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.
As a further feature of the invention cam means
are 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 t~e drill string in a generally vertical
orientation.
The valve member is preferably arranged such that
" ~L2~SS~;~
during use it is bathed in drilling fluid so that the
resulting pressure forces on the valve member
substantially balance and cancel each other out, i.e. the
valve member is essentially hydraulically neutral.
The valve structure preferably includes an
annular ring fixed to the housing and surrounding the axis
of rotation. The above-noted valve member is arranged
such that an annular flow passage is defined between
itself and this ring. The valve member is mounted for
reciprocation toward and away from the annular ring such
that the area of the annular flow passage varies from a
maximum to a minimum.
In one embodiment of the invention a reciprocal
valve member is secured against rotation while the annular
cam and cam follower are arranged to interact between the
turbine means and the valve member to effect reciprocation
of the latter on rotation of the turbine.
~ n another version of the flow pulsing apparatus,
both the turbine means and the valve member are fixed
together for both rotary and reciprocating motion. In
other words, during operation, with the cam follower in
engagement with the annular cam and with relative rotation
therebetween, the turbine and fixed valve member rotates
and at the same time reciprocate to provide for
fluctuation in the area of the annular flow passage as
described above.
By utilizing the reciprocating valve structure
described and claimed herein, a maximum restriction of the
flow area can be achieved thus enabling maximum conversion
of flow energy to dynamic pressure energy thus achieving a
maximum 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
" ~2~3555~
fluid thus increasing overall drilling efficiency.
Since the preferred form of the invention
provides a valve member that is essentially hydraulically
balanced or neutral with no substantial fluid pressure
forces thereon which would impede its movement, a highly
efficient operation can be achieved. The reciprocating
valve member is not nearly as prone to seizure by virtue
of entrapped particles and debris as compared with
relatively rotatable valving surfaces as described
previously.
The cam arrangement noted above permits timing
and frequency to be varied without being strictly
dependent on angular velocity as before and moreover tne
shape of the cam can be varied as desired to achieve the
desired shape of the pressure pulses being produced.
The above-noted annular ring portion of the valve
structure can be mounted for easy removal and replacement
from a differently sized ring thereby allowing the flow
pulsing apparatus to be tuned for a different total flow
volume.
The cam follower can be arranged to apply a non-
symmetrical force to the movable valve member thus
inducing a degree of lateral vibration which assists in
self-cleaning of the valve. However, a symmetrical
follower arrangement is also provided for when
circumstances dictate.
The reciprocating valve member, as noted above,
is resistant to the possibility of seizing due to
particles as compared with rotational valve arrangements
especially when the high degree of restriction (otherwise
known as ratio of restriction) is taken into account, i.e.
when the reciprocating valves' ability to achieve maximum
closure and maximum water hammer effect, is taken into
account. However, in the event that binding does occur,
the mechanical arrangement can be such that the valve
member simply stays open until the particle is washed
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-- 10 --
away. A degree of purposely induced vibration can assist
in cleaning. The reciprocating action of the valve
actually tends ~o push ~roublesome particles through the
valve opening and the valve surfaces can be provided with
relatively sharp edges which can assist in cutting through
certain types of particles and debris.
The operating life of the tool in terms of its
wearing ability is also enhanced by virtue of the fact
that the reciprocating valve member is essentially
hydraulically neutral and hence does no~ transfer any
resultant unbalanced hydraulic forces to the moving
assembly. Certain of the prior art arrangements, as noted
previously, are subject to large unbalanced hydraulic
forces during operation thus materially shortening their
lives.
The turbine assembly is preferably mounted on
self-cleaning sleeve bearings which have a very
substantial clearance allowing vibration of the rotary
parts relative thereto in order to induce movement of
drilling fluid into and out of these bearings. The
bearings should be made of an extremely hard material such
as tungsten carbide. The bearings are devoid of any seals
so that the drilling fluid can move freely in and out. It
has been found that this arrangement provides a relatively
long operating life and does away with the problems
associated with prior art conventional bearings which were
prone to seal damage, contamination of lubricant by grit
and rapid bearing wear.
The flow pulsing apparatus of the present
invention can be advantageously combined with a shock tool
as described hereinafter. A flow pulsing apparatus may
also be combined with an integral blade stabilizer or
reamer~ These and othe~ features including relationships
concerning the ratio of restriction to ensure pulsating
flow, relationships concerning lower and upper usable
pulsation frequencies as well as the shape of the pressure
~3S55~
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pulse will be described hereinafter.
Further features of the invention and the
advantages associated with same will be apparent ~o 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 T~E VI~WS OF DRAWINGS
Fi~-ure 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 3 is a view similar to that of Fig. 2 but
additionally incorporating a form of shock tool located
immediately below the flow pulsing apparatus;
Figure 4 is a diagrammatic view of the bottom end
of the well bore illustrating a jet of drilling fluid
emitted toward the wall and bottom of the bore hole;
Figure 5 (comprising parts 5A, 5B and 5C) is a
longitudinal half section of apparatus for producing a
pulsating flow of drilling fluid in accordance with one
embodiment of the invention;
Figures 6 and 7 are cross-section views taken
along lines 6-6 and 7-7 respectively of Fig. 5;
Figure 8 is a fragmentary half section vlew of a
flow pulsing apparatus very similar to that shown in Fig.
5 but with a modified cam arrangement;
Figure 9 is a longitudinal half section view of a
flow pulsing apparatus in accordance with another
embodiment of the invention;
Figures 10 and 11 are plan and cross-section
views respectively of typical annular cam arrangements for
the flow pulsing apparatus;
Figure 12 is an exploded view of apparatus in
~21355~L
- 12 -
accordance with the invention incorporating a shock tool
which is interposed between the drill bit and the flow
pulsing apparatus'
Figure 13 is a graph illustrating pressure
fluctuations with time;
Figure 14 is a graph illustrating the design of
the cam; and
Figure 15 is a graph relating flow area
restriction with raise in pressure Eor differing fluid
flow rates
DESCRIPTIO~ OF THE PREFERRED EMBODIMENTS
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.
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.
Flg. 1 thus demonstrates clearly the importance
of effective hole bottom cleaning ~n obtaining an adequate
drilling rate.
It is noted that Fig. 1 has been described mainly
in relation to the drilling of harder formations. In
softer formations, where the hydraDlic action of the
31213~
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.
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 ls 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 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. One of such cleaning
jets 22 is diagrammatically illustrated in Fig. 4 (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.
Fig. 3 is a view very similar to that of Fig. 2
and like co~ponents have been identified with the same
355S~
- 14 -
reference numbers as have been used in Fig. 2. However,
it will be seen from Fig. 3 that, interposed between the
flow pulsing apparatus 16, and the lower connecting sub
18, is a shock tool 24. As will be described in further
detail hereafter, this shock tool is arranged to respond
to the fluctuating or pulsing fluid flow being emitted
from flow pulsing apparatus 16 thereby to cause vibration
or oscillation of the drill bit 20 in the direction of the
drill string axis thereby to further enhance the
efficiency of the drilling operation.
Referring now to Figs. 5, 6 and 7, the flow
pulsing apparatus 16 is shown in detail. Apparatus 16
includes an external tubular housing 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. Housing 26 is in two sections which
are connected together via tapered screw threaded portion
28, with the upper end of the housing having a tapered
internally threaded portion 30 adapted for connection to a
lower end portion of the drill string. The housing 26
also includes a tapered internally threaded section 31
which may be connected to the drill bit 20 or,
alternatively, by the use of a short connecting sub, not
shown, threaded into the upper end of the shock tool 24
illustrated in Fig.3.
Housing 26 may advantageously incorporate an
integral blade stabilizer or reamer, the lobes 27 of which
are shown in phantom in Figs. 5 and 6. This enables the
IBS or reamer to be placed close to the bit without
requiring extra lengths o~ tool sections which would tend
to reduce ~somewhat or attenuate the pulsing flow and thus
reduce the efficiency of the device.
The housing 26 has a removable cartridge 32
located therein, cartridge 32 containing the turbine and
valve means to be hereafter described. For purposes of
this disclosure, the cartridge shell, which includes end
s~
- 15 -
portions 34, 36 may be considered as par~ o~ the housing
means. Cartridge shell portions 34, 36 are screwed on to
the opposing ends of a hollow metal intermediate section
38. The upstream portion 34 includes an axially arranged
nose portion 40 of outwardly stepped conical and
cylindrical shapes cen~ered in the flow passage 42 along
which the drilling fluid moves after having passed through
a screen section 44 which removes large particles
(1/8"-1/4" diameter) from the drill fluid. Screen section
44 is described in more detail in my ~,~.Patent 4,819,745-.
Upstream nose portion 40 i8 held in position by a
series of radial supports 46(Fig.6) extending bet~een such
nose portion 40 and shell portion 34.
A turbine 50 having helically curved vanes 51 to
which the fluid applies torque in known fashion and an
elongated rotor 52 is supported at its upstream end in the
nose portion 40 and at its downstream end by a turbine
stator assembly 54. Stator assembly 54 includes a
plurality of radial vanes 56 fixed to cartridge shell and
which support stator hub 58 axially in the center of the
fluid flow path.
Turbine rotor 52 includes a reduced diameter
upstream portion 60 and it i5 about this portion 60 that
an annular valve member 62 of tungsten carbide is located
for reciprocation along ~he axis of turbine rotationO
Valve member 62 co-operates with an annular valve ring 64
which is mounted in an annular recess provided in shell
intermediate section 38 and held in place by a truncated
conical entry ring 65 which bears against a shallow step
provided in the shell portion 34. Valve member 62 has a
reduced diameter portion 66 defining a throat, and a
sharply defined annular shoulder 68. An annular flow
. passage iB defined between ring 64 and annular shoulder
68, which paesage varies in area from a maximum to a
I
355~.
- 16 -
minimum as the valve member 62 reciprocates.
The nose portion 40 includes a top bearing holder
70 which supports a bearing sleeve 72 made of tungsten
carbide. Bearing sleeve 72 receives a short stub shaft
74, also of tungsten carbide, stub shaft 74 being in a
force fit relation with the upstream end of turbine rotor
portion 60. The downstream end of turbine rotor 52 also
receives a tungsten carbide bearing sleeve 76 therein in
force fit relation, which sleeve 76 receives a short stub
shaft 78, the latter being in press fit relation to a
bearing holder 80 mounted in the turbine stator assembly
54. A substantial degree of radial clearance, e.g. 0.020
to 0.050 inch, is provided between the stub shafts 74, 78
and their associated bearing sleeves so that the turbine
rotor is free to vibrate laterally during operation.
Further, since no seals are provided, the drilling fluid
is free to circulate in these relatively loose sleeve
bearings. This action sweeps away gri-tty particles which
might otherwise accumulate in the bearings and cause rapid
wear. A relatively long bearing life has been achieved in
this fashion. The lower stub shaft 78 also has a domed
end 79 which makes almost point contact with the end of
bearing sleeve 76 thus assuring low rotational drag.
The above-noted top bearing holder 70 also
supports about its outer circumference, an elongated valve
support sleeve 82 of tungsten carbide. Sleeve 82 is
suitably keyed to the bearing holder 70 and sealed thereto
with o-ring seals. The upper end of annular valve member
62 is embraced by the sleeve 82 in a relatively loose
fitting fashion, e.g. with a radial clearance of 0.020 to
0.050 inch to reduce the chances of binding due to the
presence of grit between the contacting surfaces. Valve
member 62 is restrained against rotation by means of an
axially extending key 84 (Fig. 6) fixed to bearing holder
70 and which loosely enters a slot defined in the upper
end of the valve member 62.
355~
- 17 -
In order to effect reciprocation of valve member
62 on rotation of the turbine 50, the downstream end of
the valve member 62 is provided with a cam surface 86.
Cam surface 86 is in the form of an annulus surrounding
the axis of turbine rotation. The cam shape will be
described later, it being noted here that it provides
valve opening and closing ramps, as well as dwell sections
at the valve open and valve restricted positions. In the
valve restricted position there is still enough flow as to
allow the turbine 50 to move away from the stalled
position.
The turbine rotor includes a laterally projecting
finger which acts as a cam follower 90 as it engages
annular cam surface 86. Since the valve member 62 cannot
rotate, it must reciprocate along the axis of rotation
within its support sleeve 82 if the cam follower 90 is to
remain in contact with the cam surface 86. This contact
is normally assured by two things, namely gravity, which
acts on the valve member 62, and fluid drag forces which
act on the surface of valve member 62. At the same time,
if a large piece of debris should hold the valve open
momentarily, no damage occurs as the camming surfaces
merely separate until the obstacle has been flushed away.
The use of the single cam follower finger 90
confers a special benefit in the sense that it applies a
non-symmetrical force to the valve member 62 which tends
to make it rock slightly about an axis transverse to the
reciprocation axis. This tends to provide a self-cleaning
effect, reducing the possibility of grit causing jamming
of the valve member 62 in the support sleeve 82. However,
the use of the single finger follower arrangement is not
mandatory and in Fig. 8 there is shown an identical valve
arrangement except that a symmetrical two-finger cam
follower 91 is shown in contact with the annular cam
surface 86. The vibratory cleaning effect is not present
but by using two followers, there is somewhat greater
~2~
- 18 -
design flexibility in terms of selecting the vibrational
frequency in terms of rate of turbine rotation
It should also be noted that the valve member 62
and turbine 50 are both hydraulically neutral with
hydraulic pressure forces thereon balancing and
cancelling each other out. Concerning valve member 62 it
will be noted that the drilling fluid has free access to
the interior of the member, between itself and the turbine
rotor and hence the fluid pressures can act on it in all
directions. By avoiding significant hydraulic loadings,
the contact forces at the bearings and cam surfaces are
kept to relatively low levels thus reducing wear and
helping to provide long equipment life.
An alternate embodiment of the invention is shown
in Fig. 9. As before, the cartridge 100 is in three
sections 102, 104 and 106. The upstream cartridge section
102 is provided with radial ribs 108 as before which
support a central nose portion 110. The nose portion 110
leads into an enlarged central hub 112 having an enlarged
central cavity 114 on its downstream end.
A rotor 118 extends along the axis of the
cartridge as before, turbine 116 including an elongated
rotor 118 to which is mounted a series of helical turbine
blades which respond to flow of drilling fluid by exerting
torque on the rotor 118. The downstream end of turbine
rotor is journalled in a central hub assembly 120 which is
secured by radial ribs 122. Hub assembly includes a
tungsten carbide stub shaft 124 which enters into a
tungsten carbide bearing sleeve 126 secured in the
downstream end of the rotor in a loose fit unsealed
arrangement as before.
The downstream end of rotor 118 is provided with
an enlarged portion 126 which serves to carry a pair of
diametrically opposed cam follower pins 128 both of
~2~3~55~L
tungsten carbide and secured to rotor 118 by suitable
retaining means. Follower pins 128 make contact with an
annular cam ring 130 which surrounds the axis of rotation
and which cam ring is non-rotatably mounted in an annular
recess on hub assembly 120. An annular body 132 of shock
absorbing material reduces shock loadings.
The upstream end of turbine rotor 118 is located
within the central cavity 114 on the downstream end of hub
112. Cavity 114 is provided with a tungsten carbide valve
bushing 136 held in place with retaining screws and sealed
to hub 112 by suitable O-ring seals. A sleeve-like
cylindrical valve member 138 also of tungsten carbide is
mounted to the upstream end of the turbine rotor 118 and
fixed thereto by retaining nuts 140. This valve member
138 is slidably and rotatably disposed in valve bushing
136. Hence, as the turbine is rotated by a flow of
drilling fluid, the cam action will cause the entire
turbine together with valve member 138 to reciprocate
axially up and down.
As with the previous embodiment, the central
cartridge section is provided with an annular valve ring
142 such that an annular flow passage is defined between
itself and the annular shoulder 144 defined by the valve
member 138 the area of which passage goes from a maximum
to a minimum to cause the flow to pulse as the turbine
together with the valve member both rotates and
reciprocates during operation. As before, the arrangement
is such that the valve never closes completely as there
must be at least some flow to avoid a stalled turbine
condition.
The annular cam ring 130 is shown in plan in
Fiy.10, Regions 146 correspond to the down, valve
restricted position; ramps 148 correspond to the opening
of the valve and in regions 150 the valve is full open~
Ramps 152 cause the valve to descend to the restricted
condition again.
~ 55i~i~
- 20 -
Fig. 13 is a graph of the pressure above the
restricting valve plotted against time. TCl represents
closure time while Tr represents the time the valve is
restricted. Top represents the time to open the valve
while To represents th time the valve is open. The
full cycle time is the sum of TCl + Tr + Top +
To. For best results To should be equal to or slightly
greater than the sum of the remaining times, i.e.
To ~ (TCl + Tr + Top
In Fig. 14 the cam design is represented. A
change in ramp angle A will change T~l while a change in
ramp angle B will change Top~ These angles, and the
dwell sections on the cam Tr and To are preferably
selected to satisfy the timing relationship suggested
above.
During operation 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 'nitherto 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.
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
J
s~
- 21 -
complements the pulsating flow being emitted from the
drill bit jet nozPles thereby to further enhance the
effectiveness of the drilling operation, i.e. to increase
the drilling rate.
The above-described mechanical pulsing action can
be further enhanced by the use of the apparatus
illustrated in Fig.12. In Fig. 12 a form of shock tool
160 is connected via the usual tapered screw threads 162
to the lower end, i.e. the outlet end of the flow pulsing
apparatus 16. The shock tool 160 includes an outer casing
portion 164, within which is slidably located an elongated
mandrel 166. The lower end of mandrel 166 has an
internally threaded section 168 which allows the same to
be connected to the drill bit 20 either directly or by way
of a short sub-section.
Suitable annular seals 172 and 170 are provided
between the housing 164 and the upper and lower ends of
the mandrel 166 thereby to assist in preventing
contaminants from entering between these two components
and hindering their relative axial movement. The upstream
and downstream ends of mandrel 166 are provided with a
collar portion 174 and ledge 176 and these provide annular
steps against which the upper and lower ends of a spring
stack 178 alternately engage during operation. The lower
and upper ends of spring stack 178 rest against shoulders
180, 182 respectively, fixed relative to housing 164.
This spring stack 178 is conveniently comprised of a
plurality of annular belleville-type washers although any
suitable compression spring means may be provided.
It will be seen by reference to Figure 12 that
the upper end of the mandrel, as well as the central
passageway throu~h the mandrel, which is filled with
pressurized drilling fluid during use, in effect defines
an open area piston. During operation there is of course
a pressure differential between the pressure of the
drilling fluid within the mandrel and the pressure of the
- 22 -
drilling fluid which is outside of the shock tool 160
altogether, namely, the drilling fluid which is returning
upwardly between the tool and the wall of the well bore.
By virtue of the fact that the drilling fluid leaving the
flow pulsing apparatus 16 is pulsating at a predetermined
frequency as noted above, this pressure differential also
is varying accordingly and as this pulsating differential
pressure acts on the open area piston noted above, it
serves to extend the mandrel 166 relative to the housing
164 with the result being that the shock tool 160
effectively performs as a "mud hammer". Those skilled in
this field will appreciate that for this action to take
place the drill bit weight should be reduced by lifting up
on the drill string so that the latter does not apply any
appreciable downward force to the bit. This hammering
effect is of course directly transmitted to the drill bit
20. Again, the drilling fluid leaving the jet openings 22
in the drill bit 20 will be subject to the pressure
fluctuations described above and will exhibit the desired
enhanced hydraulic effect. The shock tool 160, behaving
as a "mud hammer" applies a strong pulsing or vibrating
action to the drill bit thus causing it to drill more
effectively. At the same time, it should be realized that
the peak loadings applied to the drill bit are somewhat
less than in the case of a conventional mud hammer in
that, owing to the hydraulic action involved, the pre~sure
peaks are somewhat rounded or curved. These curved peaks
effectively do less damage to the drill bit at higher
loadings thus resulting in a longer bit life.
The use of the shock tooi 160 as shown in Fig.
12, is op~ional and under many drilling conditions its use
is unnecessary.
Although the invention is not to be strictly
limited to any particular mathematical relationship or
theory of operation, the following relationships may be
useful to those skilled in this art.
RATJO OF RESTRICTION
It was noted before that the reciprocating valve
permits a relatively high degree or ratio of restriction
of the flow to take place and consequently it can provide
a large water hammer effect. It can be shown that the
following relationship should be observed if an adequate
water hammer effect is to be achieved (neglecting drill
string elasticity):
Ar
pWcW
where:
Ao = area open to flow at the entry into the full open
valve. This area is designed in accordance with allowable
space including outside diameter of tool and mechanical
strength of the tool joint (m2).
Ar = area of the flow passage at full restriction of the
valve member, i.e. the valve member in the lowest position
(m2).
Wc = velocity of a pressure wave (sound) in drill fluid
(e.g. 1220 m/sec.)
w = velocity of the flow of drilling fluid through the
drill collars above the flow pulsing apparatus (m/s).
p = specific mass of drilling fluid, i.e. the density
(kg/m3) divided by the acceleration of gravity (m/s ).
Ho = pressure head across the open valve (Kg/m2).
FREQUENCY BOUNDARJES - (Low Frequency)
In order to avoid a frequency that would resonate
with the natural frequency of the drill collar section of
the drill string the following observations apply:
(a) For a bottom hole assembly (BHA) without the
shock tool: min frequency (f) of flow pulsing > 4212
~L~8~ii5~l
\
- 24 -
cycles/sec * L = length of the drill collar sec-tion *
based on speed of compression wave in steel of 16850
ft/sec.
(b) For bottom hole assembly (BHA) including a shock
tool (e.g. as described above): min frequency (f) of flow
pulsing ~ 1 ~ *
where:
KSt = spring constant of shock tool.
M = total mass of bottom hole assembly (slugs).
* Reference
SPE Journal Article #11228
Don W. Dareling
"Drill Collar Length is a major factor in
Vibration Control".
FREQUENCY BOUNDARIES (High Frequence)
rhe limit on high frequency requires a brief
review of the operation of the valve.
(a) when valve is fully open (Ao) - the pressure
above and below the valve is equal to a nominal pressure.
(b) valve starts to close, then becomes fully
restricted and thereafter starts to open (Ar) Pressure
above valve = Hr (head across restricted valve) +
nominal pressure. Pressure below valve = nominal pressure
r'
(c) valve opens - high pressure above valve is
released and pressure pulse moves down through valve
(Ao).
During that portion of the cycle (b) as described
above, the net downward force on the (BHA) bottom hole
assembly is increased from the normal. It is necessary
that the drill bit descend during that time interval in
order t~ funct}on efficiently. The time that it takes the
drill bit to descend is proportional to the rate of
penetration (ROP) and to the acceleration of the drill
bit.(Adb). Adb follows Newton's second law and equals
55~ ~
- 25 -
the sum of all forces acting on the bottom hole assembly
divided by the mass of the bottom hole assembly.
Jt can be shown that the time for the drill bit
(or BHA) to descend is given by the following:
Tbha = ~
where D = amount of descent/cycle (M) (related to
rate of penetration)
As noted previously, in connection with the
pressure cycle diagram To ~ ( cl r op)
and To ~ f
Tbha ~ (TC1 + Tr + Top) = f
from which it follows that
f(maximum) _ 2 x ~
An examination of the graph of Fig. 15 will
reveal some of the major advantages of the invention. The
reciprocating valve member permits the restriction area
Ar to be made relatively small. By way of example, the
rotary valve member described in my ~k~. Patent No.
~,a~ 9
1,815,7~, by virtue of the required radial
clearances and the required size of the valve ports, was
not able to provide a restriction area of less than about
0.60 square inchesO However, with the reciprocating valve
of the present structure, the res~riction area (Ar) can
be made as small as desired just so long as sufficient
flow can be provided as to move the turbine away from the
talled condition. The water hammer effect increases at a
very high rate as the restriction area decreases,
especially in the areas when the slopes of the curves have
decreased and ~ake shallow angles with the horizontal
(pressure) axis. The effect is especially notable at low
total flow rates. With previous rotary designs and a
maximum permissible restriction of about 0.60 Rquare
inches, the water hammer effect only provides a pressure
~Q'
~ i;Si5~
- 26 -
rise of about 100 psi at a flow rate of 230
gallon/minute. With an area restriction (Ar) of 0.20
square inches at the same flow rate, the pressure rise
when the valve is restricted is over 1000 psi, a ten-fold
difference. The effect is somewha~ less dramatic at
higher flow rates but in all cases the increase in water
hammer effect coupled with the greater flow rates made
possible by the larger valve open area (Ao) provide a
very effective flow pulsing operation and enable higher
drilling rates to be achieved than hitherto.
