Note: Descriptions are shown in the official language in which they were submitted.
2~35
This invention relates to shock subs for use in drill strings
and methods of using the same.
It is well known in the art of drilling wells that the drill
string and bit are subjected to substantial axial oscillations. As
discussed in Down Hole Measurements of Drill String Forces and
Motions, Journal of Engineering for Industry, May 1968, pages 217-
225; Analysis of Down Hole Measurements of Drill String Forces and
Motions, AIME Transactions, May 1968, pages 208-216; and Longitudi-
~nal and Angular Drill String Vibrations with Damping, Journal of
Engineering for Industry, November 1968, pages 671-679, these ',
oscillations are caused by pressure variations in the drill string,,
impact of bit teeth against the rock face being drilled, bit cone
'action against the rock face being drilled and rotation of the drill
string. Under some circumstances, the oscillations can be so great
as to cause the bit to move vertically off the rock face being
drilled. Even in situations where the oscillations are not so
severe, the oscillations can cause decreased bit tooth life,
shortened bit bearing life, decreased fatigue life of the drill
collars and drill~ pipe and shortened fatigue life of the derrick. I
It is accordingly not surprising that considerable effort has been !
spent on vibration damping devices for incorporation in a drill
string.
Broadly, drill string shock subs comprise an outer barrel
having a screw connection at one end, a mandrel tel~scopingly
received in the barrel and having a screw connection at the opposit~
end, and a spring mechanism of some description operating between
the barrel and mandrel. Spring mechanisms of wide variety have beer
suggested for use in shock subs. Exemplary spring designs include
annular abutted metallic plates as shown in United States patent
2,570,577; helical springs as shown in United States patent 3,122,90l2;
- 2 -
i
~5~3235
,gas springs as shown in United States patent 3,230,740; pivoting
metal sections as shown in United States patents 3,254,508 and
3,447,340; Belleville springs as shown in United States patent
3,871,193; and rubber or rubber-like springs as shown in United
States patents 3,033,011; 3,099,918; 3,301,009 and 3,339,380.
The spring types in commercially available shock subs
correspond broadly to the spring types illustrated in the patents
cited above to the extent that there are available shock subs having
gas springs, Belleville springs, rubber springs and helical springs.
Each major type spring has substantial disadvantages. The gas
,springs exhibit a spring rate that varies in response to drilling
depth as do some of the ru~er spring devices. Although other of
the rubber spring devices exhibit a constant spring rate with
respeCt to depth, the spring rates are quite high, usually in the
range of 90,~0C-120,000 pounds per inch of deflection. Although
the hel.ical springs exhibit a relatively low constant spring rate,
e.g. 35,000 pounds per inch of deflection, this type tool is quite
long which tends to destabilize the drill string. In addition, the~
helical spring is machined from a single piece of steel stock and ig
accordingly inherently expensive.
There are three interrelated design criteria which dictate
the performance of a shock sub: load capacity, spring rate and
deflection. It is suggested in the prior art that the lower the
spring rate, the more effective will be the shock sub in damping
vibrations as pointed out in the article entitled Longitudinal and
Angular Drill String Vibrations with Damping, supra" and a publica-
tion of Johnston-Schlumberger, entitled "Shock Guard Drilling
Shock Absorber." The load capacity of a shock sub is normally
'dictated by the maximum weight applied to the bit during drilling
,which is normally a function of bit diameter. Although the maximum
. - 3 -
~n23s
weight applied to drillinq bits varies somewhat, it is normally in
the range of 6,000-9,000 pounds per inch of bit diameter. Load,
'deflection and spring rate in a mechanical spring assembly are
irelated by the equation:
1,; . I
Load = _ Deflection . (1)
Spring Rate
Accordingly, when the load capacity is dictated by external circum-¦
;
;stances, one can only decrease the spring rate of the system by
- 'increasing deflection. '
Another design constraint to be reckoned with is the maximum !
, ~
diameter available for the spring. Because the nominal hole size i~
.. ' i
dictated by the size of the bit, it is self-defeating to contemplatei
utilizing a larger bit to give more tool area for two reasons.
First, the industry will dictate the size bits to be used and the
tool must be designed accordingly. Second, the larger the bit, thej
greater the load capacity required. There is, however, very little
theoretical design constraint imposed on the axial length of the
spring although it is desirable that the shock sub be as short as
possible to maintain the drill collars as stable as possible.
To illustrate the problem, a typical bit size is 7 7/8" in
diameter and accordingly has a maximum design weight applied theretq
in the range of 47,000 - 70,000 pounds. Because there must be
sufficient clearance between the external surface of the shock sub
and the wall of the hole to allow cuttings and mud to pass there-
through, the maximum external diameter of the shock sub can be no
greater than about 6 7/8". The mandrel must have a conduit providin g
a passage for mud traveling downwardly to the bit of about the same
size as the passage through the drill pipe and drill collars. In
~addition, the passage through the mandrel must be of sufficient size
-- 4 --
~5q;~Z3S
to receive a fishing spear. Accordingly, a passage in the range of~
l 1~2" - 3" is required. It will accordingly be seen, using average
values, that the barrel, mandrel and spring can occupy an annular
area having an external diameter of about 6 7/8" and an internal
diameter of about 2 1/4". Because a great deal of this area must be
occupied by structural steel members comprising the wall of the
,barrel and wall of the mandrel, there is precious little design
freedom in selecting the operating diameters of the spring mechanism
installed between the mandrel and barrel.
Although the amount of weight applied to the bit is a functiqn
of diameter, normally the amount of weight and accordingly the load
capacity of the spring assembly of this invention is in the range
of 10,000 - 125,~00 pounds. With more typically sized bits, the
amoun~ of weight and the load capacity is typically in the range of
30,000 - 75,000.
To further illustrate the effects of design constraint on
the spring assemblies, it is believed that the 35,000 pound/inch
spring rate of the helical spring in the shock sub offered by
Johnston-Schlumberger probably represents a design which minimizes
spring rate as far as practicable for the load capacity required of
shock subs. In other words, it is believed impossible to design a
practical helical spring having a substantially lower spring rate
and providing the requisite load capacity and life expectancy withi~
the constraints dictated by hole size, central passage size and
tool length.
The shock sub of this invention is broadly organized to
include an outer barrel having a screw connection on one end for
connection to either the bit or the drill collars, a mandrel
~tele~copingly received in the barrel and having a screw connection
on the other end for connection in the drill string, and a spring
assembly operating between the barrel and mandrel to damp relative
movement between the barrel and mandrel. The snock sub is accordin ly
reversible in the sense that it can be run with the barrel connectia n
~t~35
down or with the mandrel connection down. As will be more fully
pointed out hereinafter, a given shock sub of this invention has
significantly different load capacities, because of the application
of hydrostatic pressure inside the tool, when reversed.
The spring assembly in the shock sub is designed to provide
a load capacity of 10,000 - 125,000 pounds and more typically has
a load capacity of 30,000 - 75,000 pounds. The deflection of the
!
spring is substantially less than twenty inches and the spring rate
is less than 25,000 pounds per inch of longitudinal tool movement
and is more desirably less than 15,000 pounds per inch of longitu-
dinal tool movement.
The spring assembly com~rises a multiplicity of annular inne~
springs having radially outer surfaces inclined to the axis of the ¦
shock-sub and a multiplicity of annular outer springs having radial~y
inner surfaces inclined to the axis of the shock sub. The inner an~
outer springs are stacked alternately in a chamber between the
~arrel and mandrel so that collapsing movement of the barrel and
mandrel causes the springs to move axially toward each other. The
inclined surfaces of the inner and outer springs engage during
collapsing movement of the barrel and mandrel causing the outer
springs to be stressed in tension and the inner springs to be
stressed in compression substantially only in a multiplicity of
planes perpendicular to the axis of the shock sub. The spring
assembly of this invention accordingly utilizes what are sometimes
called "ring springs" which are broadly known in the art as shown in
United States patents 745,425; 1,598,228; 1,689,662; 1,700,133;
2,413,740; 2,515,346; 3,073,585; and 3,536,314 as well as in
Characteristics of Ring Springs by Tyler G. Hicks, American
Machinist, 1928, pages 192-195 and Mechanical Springs, First Edition ,
Arthur M. Wahl, Penton Publishing Co., 1944, pages 348-358.
-- 6 --
1~5~Z35
In summary of the above, therefore, the present inven-
tion may be considered as a shock sub for connection in a
drill string, comprising a barrel having means at one end for
connection to the drill string; a mandrel mounted in the
barrel for telescoping movement along a longitudinal axis and
having means at one end for connection to the drill string;
means for transmitting torque between the mandrel and barrel;
a ring spring assembly for damping relative movemen-t between
the mandrel and barrel, the assembly comprising a multiplicity
of inner and outer annular ring springs residing in planes
transverse to the axis, the inner ring springs comprising
radially outer surfaces inclined to the axis, the outer ring
springs comprising radially inner surfaces inclined to the axis
and engaging the radiallv outer surfaces of the inner ring
springs; means carried by the mandrel and barrel for
expanding the outer ring springs and compressing the inner
ring springs upon collapsing movement of the mandrel and
barrel; the ring springs reaching the yield point of the
material thereof at a predetermined amount of collapsing
movement of the barrel and mandrel; and means carried by the
barrel and mandrel for limiting the amount of collapsing
movement therebetween to less than the predetermined amount.
Other features and a fuller understanding of the
invention may be had by reference to the following description
taken in conjunction with the accompanying drawings and claims.
IN THE DRAWINGS-
Figure 1 is a sectional view through the earth illus-
trating the drilling of a hole therein by a drill string
incorporating a shock sub of this invention;
Figures 2A and 2B are partial longitudinal vertical
cross-sectional views of the shock sub of this invention;
~15~Z35
Figure 3 is an exploded broken isometric view of the
ring springs incorporated in the shock sub of Figures 2A
and 2B;
Figure 4, appearing on the same sheet as Figure 1, is
a partial vertical cross-sectional view of the device of
Figures 2A and 2B illustrating the replacement of a plurality
of ring springs with an insert;
Figure 5, appearing on the same sheet as Figure 1, is
a view similar to Figure 4 illustrating another technique
for stressing the ring springs;
Figure 6 is a load-deflection diagram of the tool of
Figures 2A and 2B; and
Figure 7, appearing on the same sheet as Figure 1, is
a partial vertical cross-sectional view of another embodiment
of the invention.
3`23~ 1
Referring to Figure 1, there is illustrated a hole 10 being
drilled in the earth through a string 12 of surfaee pipe placed
therein and bonded to the earth by a eement sheath 14. The hole 10
is being drilled with a bit 16 connected by a drill string 18
suspended from a travelling block 20 by a conventional hook up 22
ineluding a kelly 24. The travelling bloek 20 is suspended from a
suitable crown bloek (not shown) provided by a suitable drilling
rig (not shown). Drilling fluid is p~mped through a mud line 26 in~o
the drill string 18, exits through nozzles (not shown) in the bit
16 and then circulates upwardly through the annulus between the holq
10 and the drill string 18 to remove euttings, eool and lubrieate
the bit 16, and control formation pressures as is customary in the
art. I
- The drill string 18 eomprises a shoek sub 28 of this inventiqn,
a multiplicity of drill eollars 30 and a string of drill pipe 32
illustrated as ineluding a number of pipe joints having externally
upset box and pin connections which are universally screw thread
eonneetions. The function of the drill string 18 is to conduct
drilling fluid to the bit 16, to transmit torque to the bit 16, to
stabilize the direction oL drilling, to provide means for removing
the bit 16 from the hole 10, and to apply weight to the bit 16.
More speeiically, the general funetion of the drill pipe 32 is to
provide a mechanical and hyaraulie eonneetion to the drill collars
30. The drill collars 30 provide a mechanical and hydraulic connec~
tion between the drill pipe 32 and the bit 16 but also act to
apply weight to the bit 16 and to stabilize the direetion of drilling.
The drill co1lars 30 are typically massive pipe joints providing a
substantial amount of weight immediately above the bit 16 and are as
inflexlble as praeticalities allow.
1.
- 8 - I
.3235
One of the early lessons in drilling with rotar~v drill pipe I
was that the bulk of the drill string 18 must be kept in tension wi~h
only the lower part of the drill collars allowed to be in compressign
in order to drill a relatively straight hole. Thus, the so-called
neutral point 34 divides the drill string 18 into a relatively
short lower section which is in compression and a relatively long
upper section which is in tension. Accordingly, if it is desired tc
maintain 50,000 pounds of weight on the bit 16, it would not be
unusual for the total weight of the drill collars 30 to be in excesC
of 100,000 pounds in order to compensate for the bouyancy of the
drill collars 30 in the drilling fluid, to accomodate the reaction
force of mud passing through the bit nozzles, and to maintain the
neutral point 34 far below the top of the drill collars 30.
_ As pointed out previously and as discussed in great detail
in the articles in the Journal of Engineering for Industry and in
the AIME Transactions cited above, the drill string 18 is subjected
to substantial axial oscillations during drilling. The data indi-
cates that the provision of shock subs can substantially decrease
the maximum amplltude of the oscillations, particularly at relative y
high rotary speeds. In addition, the data indicates that a shock
sub having a spring rate of 500,000 pounds per foot of deflection
(41,668 pounds/inch of deflection) is as effective at low rotary
speeds as a shock sub having a spring rate of 1,000,000 pounds per ¦
foot of deflection (83,333 pounds/inch of deflection) and is
considerably more effective at higher rotary speeds. ~3ecause high
rotary speeds are particularly desirable to achieve high penetratior I _
rates when drilling relatively hard formations, it is believed that !
shock subs having even lower spring rates would be more desirable
at high rotary speeds. Although the present suppliers of shock
subs have obviously taken substantially different views of the
importance of low spring rates, Johnston-Schlumberger is quite prou~
g
~q3Z35
of the spring rate of 35,000 pounds per inch of deflection in its
helical spring type shock sub. It is accordingly apparent that
low spring rate shock subs are deemed to be highly desirable by
some segments of the industry.
Referring to Figures 2A and 2B, there is illustrated the
shock sub 28 of this invention. Broadly, the shock sub 28 comprises
a barrel 36, a mandrel 38 telescopingly received in the barrel 36,
means 40 for transmitting torque between the barrel 36 and mandrel
38, a spring assembly 42 for damping relative movement between the
barrel 36 and mandrel 38, and means 44 for equalizing the fluid
pressure around the spring assembly 42 with that inside the mandrel
38.
The barrel 36 may be of any suitable construction and is
illust-rated as comprising a plurality of connected threaded compo-
nents including an upper body or washpipe barrel 46 having a female
threaded connection 48 for attachment in the drill string 18, a
central passage including a relatively small initial section 50 and
an enlarged lower section 52. The lower end of the upper body 46
terminates in a threaded end 54. For purposes more fully apparent
hereinafter, the upper body 46 also includes one or more threaded
openings 56 transverse to the tool axis 58 which is closed by a
threaded sealing plug 60.
The next section of the barrel 36 is a spline body 62 having
an upper end 64 providing male threads engaged with the threaded end
54 of the upper body 46. The spline body 62 provides a central
passage 66 sized to closely receive the mandrel 38 and provides a
plurality of circumferentially spaced longitudinally or helically
extending splines 68 extending interiorlyiof the passage 66. The
spline body 62 terminates in a lower end 70 having male threads
thereon.
-- 10 --
:1~5~3235
The next section of the barrel 36 is a middle body 72 having .
an upper threaded end 74 engaged with the threads of the lower end .
70 of the spline body 62. The middle body 72 provides a central
passage 76 of substantially greater diameter than the passage 66 and
one or more transversely extending passage 78 having a threaded
sealing plug 80 therein. The body 72 terminates in a lower threaded
end 82.
The barrel 36 terminates in a packing nut 84 having an upper
threaded end 86 engaged with the threads of the middle body 72. The
packing nut 84 provides an internal passage 88 sized to closely
receive the mandrel 38 and is intermediate in diameter to the pas-
sages 66, 76. The lower terminal end of the packing nut 84 provides
a piurality of recesses each of which receives an annular packing
member- 90. A surface 92 perpendicular to the tool axis 58 consti-
tutes the end of the packing nut 84.
The mandrel 38 includes an enlarged lower end 94 having a
female'threaded connection 96 in the end thereof for attachment to
the drill string 18. Because the shock sub 28 has female connections
at its opposlte ends, it can be connected between the bit 16 and
lowermost drill collar 30 without requiring the use of an adapter
sub. The lo~er end 94 terminates in a shoulder 93 facing the
barrel 35 which is perpendicular to the tool axis 58. It will be
seen that the maximum travel between the barrel 36 and mandrel 38,
in a collapsing or telescoping direction, is dictated by the spacing
between the surface 92 and the shoulder 98.
The mandrel 38 also includes a mandrel body section 100
integral with the enlarged end 94 and includes a cylindrical exter-
nal surface 102. One or more grooves 104 are cut in the surface 102
adjacent the packing nut 84 which receive a wear ring 106. The wear
rings 106 are typically made of an organic polymeric material which
is subject to being transferred to an adjacent metal surface by
-- 11 --
23S
friction contact and which exhibits a low coe~ficient of friction.
,One exemplary material for the wear rings 106 is tetrafluoroethylenq.
During relative movement of the mandrel 38 and barrel 36, the materil~
ia~ of the wear rings 106 is distributed onto the surface of the
passage 88 and thereby acts to minimize or preventlscuffing of the
,~surface 102 and the passage 88.
The body section 100 joins another mandrel body section 108
of reduced diameter having an external surface 110. The diameter o
the cylindrical surface 110 is only slightly less than the diameter,
of the passage 66 of the spline body 62. It will be seen that there
is provided an annular chamber-between the internal passage 76 of
the middle body 72 and the surface 110. Adjacent a shoulder 112,
the mandrel body section 108 provides an exteriorly threaded sectio~
114 which receives a mandrel nut 116 having female threads 118
engaging the threaded section 114. A set screw 120 extends through¦
~the mandrel nut 116 and acts to secure it in position.
The external surface 110 of the body section 108 provides,
adjacent the spline body 62, a plurality of grooves 122 which
are preferably longitudinally extending but which may be of helical~,
shape, receiving the splines 68 of the spline body 62. It will be '
evident that the splines 68 and grooves 122 cooperate to provide the
,torque transmitting means 40 so that rotation of the barrel 36
results in rotation of the mandrel 38. Il
The upper extremity of the mandrel body section 108 includes ¦
a threaded end 124 receiving an internally threaded end 126 of a
washpipe 128 comprising part o~ the mandrel 38. The lower end of
the washpipe 128 includes a surface 130 perpendicular to the tool
axiS 58 and cooperates with a facing perpendicular surface 132 '
provided by the spline body 62 to limit expanding or untelescoping
movement of the barre1 36 and mandrel 38.
- 12 -
~s li
A central passage 134 extends through the mandrel 38 from the
top of the washpipe 128 through the threaded connection 96 and allows
the delivery of drilling Eluid from the passage sections 50, 52
through the end of the shock sub 28. As suggested previously, the
passage 134 should be no smaller than about 1 1/2" in diameter to
allow the passage of drilling fluid without undue pressure loss and
to allow a fishing spear to enter the washpipe 128 in the event the
mandrel 38 becomes lost in the hole 10.
Referring to Figures 2B and 3, the spring assembly 42 includes
a plurality of inner annular ring springs 136 and a plurality of
outer annular ring springs 138. The ring springs l.36, 138 are
alternately stacked in the chamber provided between the barrel 36
and'the mandrel 38. As will be seen, telescoping or collapsing
.movement of the barrel 36 and mandrel 38 causes each of the springs
136, 138 to be stressed substantially only in a plane perpendicular~
to the tool axis 58. When stressed, the inner ring springs 136 are
placed in compression while the outer ring springs 138 are placed
in tension.
All of the inner ring springs 136 are desirably identical..
Although the cross-sectional shape of the rings springs 136 may
vary substantially as pointed out in the publications and patents
relative to ring sprin~s cited previously, in order to maximize the,
cross-sectional area of the inner ring springs 136 while maximizing
the allowable external diameter of the body section 108, the, inner
ring sprin~s 136 preferably comprise a radially inner cylindrical
surface 140. The inner diameter of the surface 140, in either the I ~
unstressed or fully stressed condition of the ring springs 136, is
larger than the diameter of the cylindrical surface 110. As will
be more fully pointed out hereinafter, the maximum stress applied
to the ring spring assembly 42 and consequently the maximum stress
- 13 -
~5~35
applied to each of the springs ]36 is controlled by the amount of
.
collapsing movement allowed by the spacing between the surfaces 92,
98.
The inner ring springs 136 also include upper and lower
surfaces 142, 144 which are generally transverse to the tool axis
58 and which are preferably perpendicular thereto. The radially
outer surface of the inner ring springs 136 includes a pair of
surfaces 146, 148 which are frustoconical in configuration and ~hich,
include a common maximum diameter. The surfaces 146, 148 define
oppositely facing acute angles 150, 152 respectively with the tool
axis 58. As will be more fully apparent hereinafter, it is preferred
that the angles 150, 152 be substantially ïdentical.
The outer ring springs 138 are also desirably identical.
Although the outer ring springs 138 may also be of any suitable
shape, in order to maximize the cross-sectional area in a plane
includinq the tool axis 58, the ring springs 138 include a radially
outer cylindrical surface 154. The diameter of the radially outer
surface is less, in either the unstressed or fully stressed condition
of the spring assembly 42, than the diameter of the passage 76
provided by the middle body 72.
The outer ring springs 138 include upper and lower surfaces
156, 158 which are transverse to the tool axis 58 and which are
preferably perpendicular thereto. The radially inner surface of the
outer riny springs 138 include a pair of surfaces 160, 162 which are
frustoconical in configuration and which include a common minimum
diameter. The surfaces 160, 162 define oppositely facing acute
angles 164, 166 respectively with the tool axis 58.
If the angles 150, 152, 164, 166 are substantially identical,
this will allow all of the inner ring springs 136 to be identical
and all of the outer ring springs 138 to be identical. This is, of
- 14 -
32~
course, highly desirable when it is necessary to replace any of thel
springs 136, 138. In addition, this provides considerable simplici~y
.in design and manufacture of the shock sub 28.
` The design selection of the quantity for the angles 150, 152,
164, 166 is of substantial importance as is the cross-sectional area
of the inner and outer ring springs 136, 138. The stress induced i~
,the springs may be calculated from the equation
~, .
S = (Load) tan a ( 2)
where S is the maximum stress in each spring, A is the cross-sectio~al
area of each individual spring, ~ is the angle 150, 152, 164, 166
.and K is a constant of the system and is a function of a. Thus, in
order-to provide for the maximum load induced in each spring, the
cross-sectional area A should be maximized. In addition., the
cross-sectional areas of the springs.136, 138 are desirably identi-¦
cal. ~he value of K may be obtained from the equation provided in ¦
the publication of Wahl cited previously. The maximum value for
the angles 150, 152, 164, 166 can be obtained from equation (2)
because the maximum value of S is a function of the material of the .
springs 136, 138 and the load value is dictated by the desired
capacity of the shock sub 28. It may be, however, that one may wisl
to select a smaller value for a because the spring rate of each
individual spring will generally decrease as the value of ~ decreas~ s.
In generaL, the value of a in shock subs of this invention will lie
in the range of 5-20 and a preferred range is on the order of 11-1 O
One of the oddities of ring spring assemblies lies in the
relationship between the number of springs and the spring rate of
''the assembly. Because the weight applied by the drill collars 30 tc ,
the barrel 36 is transmitted through each of the ring springs 136,
138, it will be seen that each spring is subjected to the entire
Z35
applied load. Accordingly,
i Load
SRa = - (3)
(~s) (n)
.' I
l where SRa is the spring rate of the assembiy, Ds is the deflection ¦
,lof each individual spring and n is the total number of springs.
Because the design load remains the same, the deflection of each
spring in an axial direction will remain the same, independently ~!
the number of springs employed, because the amount of each deflecti~n
of each spring is directly related to the applied load. Thus, the ¦
' sprincf rate of the assembly 42 may be modified merely by changing
the ~umber of springs in the assembly 42. This has two significant
aspects. First, in the design of this shock sub the desired
spring rate can be achieved merely by selecting the number of
springs. Second, the spring rate of an existing shock sub can be
increased by replacing some of the springs with a spacer or can
be decreased by removing a spacer from the spring chamber and
replacing it with operative springs. .
In this regard, Figure 4 illustrate`s an annular spacer 168
of rectangular cross-section positioned between the lower face 170
of a spline body 172. A ring spring assembly 172 is provided
with a half spring 174 abutting the spacer 168 and engaglng the
first full spring 176 of the assembly 172. It will be evident
that a shock sub of this invention may be designed to normally
incorporate the spacer 168 so that the spring rate of the tool may
¦ be decreased merely by removing the spacer 168 and half spring 174,
placing a number of operative full springs in the spring chamber
'and replacing the half spring 174. In the alternative, the spring ¦
rate of the sub 28 may be increased by replacing several of the
springs 136, 138 with the spacer 168 and half spring 174.
- lG -
Il
3Z35
There are a variety of suitable techniques for stressing the
assembly 42 in response to telescoping movement of the barrel 36 and
mandrel 38 in addition to the showing of Figure 4. One exemplary
technique is illustrated in Figures 2A and 2B where a first projec-
tion 178 is integral with the lower end 70 of the spline body 62 and
provides an inclined surface 180 of the same angle as the surface
162. The surface 180 engages the radially outer surface 146 of the
uppermost inner ring spring 136. The projection 168 also includes a
terminal face 182 perpendicular to the tool axis 58 and is comparable
to the surface 158. The mandrel nut 116 also provides a projection
184 having an inclined surface 186 comparable to the surface 160 and
engages the surface 148 of the lowermost ring spring 136. The
projection 184 provides an end surface 188 perpendicular to the tool
axis 58.
Another exemplary technique is illustrated in Figure 5 where
a spline body 190 includes a lower face 192 perpendicular to the tool
axis and engaging a half spring 194 providing an inclined surface
196 engaging the appropriate inclined surface 198 of the uppermost
full spring 200. Although not illustrated, the mandrel nut facing
the spline ~ody may be of similar configuration abutting a similar
half spring. It will be noted that the showings of Figure 4 and
5 are substanti.ally identical except for the provision of the spacer
168.
Referring back to Figure 2A, the pressure equalizing means
44 includes a floater 204 of annular construction having an internal~
diameter sized to slide close.ly on the external surface of the
washpipe 128 and an external diameter sized to slide closely on the
wall of the l~assage section 52. The floater 204 is preferably made
of organic polymeric material, such as Ryton, and provides one or
more interior annular grooves receiving a packing member 206 and one
or more external annular grooves receiving a packing member 208. A
floater stop 210 provides one limit of sliding movement of the
- 17 -
~5~3s
floater 204 and a shoulder 212 on the washpipe 128 provides another
limit of travel.
The function of the floater 204 is to equalize the pressure
in the spring chamber with the pressure inside the shock sub 28.
This is accomplished because the floater 204 moves downwardly upon
increase in hydrostatic pressure inside the mandrel to increase the
pressure in the spring chamber or moves upwardly when the hydroqta-
tic pressure inside the mandrel declines. This has two effects.
~irst, the pressure differentials sensed by the seals 206, 208
and the packing members 90 will be substantially reduced. Second,
there would be a tendency for the mandrel 38 and barrel 36 to
lock up hydraulically because the spring chamber is filled with
a lubricant and because the spring chamber is variable in size
depend-ing on the relative position of the barrel 36 and mandrel 38.1
It will be seen that the floater 204 resolves these difficulties in j
a simple and expeditious manner.
In order to fill the spring chamber with a lubricant, one of¦
the sealing plugs 60, 80 may be removed and a liquid lubricant
poured therethrough. In order to prevent the accumulation of air i~
the lubricant filled annulus between the mandrel 38 and barrel 36,
both sealing plugs 60, 80 may be removed, lubricant pumped into the
lower of the openings 56, 78 while allowing air to discharge from
the other opening and periodically tilting the sub 28.
One of the features of the shock sub 28 which is not immedi-¦
ately apparent is that the max-mum deflection allowed by the spacinc~
between the shoulders 92, 98 is selected to be related to the
~maximum permissible deflection of the spring assembly 42 which is
related to the maximum permissible deflection of each of the ring
springs 13~, 138. The maximum permissible deflection of each
springi136, 138 may be calculated from the teachings of Hicks and
- 18 -
235
~ahl, supra, and is a function oE the elastic limit of the material
selected. The maximum deflection of the spring assembly 42 is, of
course, the maximum deflection of each spring multiplied by the
number of springs. This distance is then selected for the spacing
~etween the shoulders 92 ! 98. Thus, the spacing between the shoul-
~èrs 92, 98 is selected so that the mandrel 38 and barrel 36 bottom
put before the elastic limit of the springs is reached.
When the shock sub 28 is utilized in the drill string 18, the
bit 16 is desirably threaded into the screw connection 96 of the
mandrel 38 and the lowermost drill collar 30 is threaded into the
screw connection 48 of the barrel 36. When drilling mud, for
example, is pumped down the drill string 18, there is created a pump;
~open force F acting on the mandrel 38.~hich tends to untelescope the
mandrel 38 and barrel 36. The amount of the pump open force F is
readily calculated by multiplying the unbalanced areas of -the
mandrel 38 by the pressure acting thereon. Another force acting on
the shock sub 28 during drilling is a reaction force Fr, usually
called bit thrust, caused by passage of the drilling fluid through
the nozzles of the bit 16. In order to stress the spring assembly
42, the weight applied by the drill collars 30 must exceed the pumpl
open force F less the bit thrust Fr. Normally, the pump open force ¦
F is greater than the bit thrust Fr. Thus, the spring assembly 42
is not stressed until a sig^nificant amount of weight is placed on
the barrel 36. The difference between the pump open force F and the
bit thrust Fr can be considered to be a base line or threshold in
the sense that the weight of the drill collars equal to this value
must be applied to the shock sub 28 before the damping characteris- i
tics of the spring assembly 42 are brought into play.
As sufficient weight is applied to -the barrel 36, the ring
springs 136, 138 are stressed to transmit the applied load to the
.
- 19 -
,
!
~L~L~235
mandrel 38 and consequently to the bit 16. As the bit 16 is rotated
on the bottom of the hole 10, the oscillations induced in the mandrel
38 are damped during transmission through the spring assembly 42 to
the barrel 36.
A shock sub in accordance with the principles of this inven-
tion has been constructed to acco~ate a nominal load of 50,000
pounds and has a maximum deflection of 4 5/8". The spring assembly
42 comprises twenty inner ring springs 136 and twenty outer ring
springs 138. The unstresséd height of the spring assembly 42 is
about twenty five inches. The tool is 6 7/8" in outer diameter and
provides an internal passage 134 of 1 3/4" in diameter. The sub is
about eight feet long.
One of the advantageous features of the shock sub 28 is its
short~length. It will be apparent that the stability of the lower
end of the drill string 18 is promoted by a short shock sub. Thus,
the shock subs of this invention are characteriz~ed by a relatively
short length, usually under ten feet long and preferably about eight
feet long.
The constructed tool was placed in a jar testing press so
that various loads could be applied thereto and the displacement of
the barrel 36 relative to the mandrel 38 measured. In the test
setup employed, no hydrostatic pressure was applied to the interior
of the mandrel 38 so that the ?ump open force F was not created nor
was drilling fluid discharged through a bit nozzle so that the bit
thrust Fr was not created.
In theory, the load applied to the barrel 36 should immedia-;
tely cause relative movement between the barrel 36 and mandrel 38.
As a matter of practicality, there may be a certain amount of fric-
tion generated in the spring assembly 42 and there may be required
a minimum application force Fm to center the ring springs 136, 138
- 20 -
Z35
and cQmmence stressing thereof, at least during initial or break in
runs of the sub 28. Part of this minimum application force Fm is
required to overcome friction and part is a result of manufacturing
tolerances. The spriny rate of the constructed tool was designed to
be 12,500 pounds per inch of deflection. Data obtained from the
first test generated a load-deflection curve 214 which was substan-
tially linear over the test range of 0 - 50,000 pounds as shown by
comparison with the line 216 even though the spring rate differed l
slightly from the design value. A second test, incorporating improved
deflection measuring techniques, generated a load-deflection curve
218 which compared very favorably with the line 220 representing a ~
spring rate of 12,500 pounds per inch of deflection. It is not yet ¦
clear whether the difference between the curves 214, 218 is a functipn
of improved measurement during the tests or whether it relates to a
breaking in of the shock sub.
It will be evident that the spring rate of the constructed
tool, which is the slope of the line 218, was remarkably constant
throughout the load test range of 0 -50,000 pounds. Accordingly,
when used in a drill string, the spring assembly will begin to
stressed at a load value of Fm + F - F and will thereafter exhibit
a load deflection curve 222. It is conceivable, of course, that
Fm is quite small as suggested by the curve 218.
One of the peculiarities of the shock sub 28 is tha~ it may
be attached in the drill string 18 with the mandrel end 94 facing
upwardly rather than downwardly. The effect of this bidirectionalit~
of the shock sub 28 requires some discussion of pump open forces
,generated during drilling. It is a common teaching in the art that ¦
the amount of the pump open force is determined by the pressure in
'a shock sub which acts on unbalanced areas and tends to untelescope
the tool. In shock subs of the prior art and in the shock sub 28
when the mandrel end 94 is run down, i.e. facing toward the bit, an
- 21 -
/
!
z35
amount of weight from the drill collars equal to the pump open force
must be applied to the barrel in order to commence stressing of the
spring assembly.
When the shock sub 28 is inverted, i.e. the mandrel end 94
faces the drill collars 30, the direction of the pump open force is
changed and a much smaller amount, if any, of weight from the drill
collars 30 must be applied to the mandrel in order to commence
stressing of the spring assembly 42. In this situation, the pump
open force is acting in a direction opposite to the bit thrust Fr
rather than in the same direction. It is conceivable, of course,
that with the shock sub 28 inverted, the pump open force could be
nil or zero depending on the amount of unbalanced areas inside the
shock sub 28 when it is inverted. With the sub 28 inverted, the
threshold force tending to stress the spring assembly 42 is reduced.
This is of considerable importance at any time the load to be applied
to the bit 16 is less than about half the load capacity of the spring
assembly 42 when the shock sub 23 is normally oriented. In other
words, a lower amount of weight applied by the drill collars 30
begins stressing the spring assembly 42 with the mandrel end 94 up
than with the mandrel end ~4 down: An exemplary situation where
this is of value is in drilling at very shallow depths where only a ¦
few drill collars 30 can be placed in the hole 10. In this situatio~,
very little weight can be applied to the bit 16. Accordingly, it is
advantageous to drill the surface hole with the sub 28 inverted to
take advantage of the lower spring threshold and increased sensi- ¦
tivity and then drill the remainder of the hole with the mandrel end
94 down as shown in Figure 1. Other circumstances where it is
.
desirable to invert the sub 28 will be recognizable by those skilled¦
in the art.
Referring to Figure 7, there is illustrated another embodiment~
of the inven~ion comprising a ring spring assembly 224 including
inner and outer ring springs 226, 228 and a mandrel nut 230 secured
- 22 -
235
in a unit handled mass by an annular sleeve 232 of permeable mater-
ial such as nylon net or permeable butyl rubber. Inside the sleeve
232, the individual ring springs 226, 228 define gaps therebetween
which are substantially filled by a porous and permeable foam
material 234. The sleeve 232 and material 234 are sufficiently
permeable to pass lubricant toward and away from the springs 226,
228. It will be seen that the springs 226, 228 are capable of
being unit handled along with the mandrel nut 230 if desired.
Although this invention has been described in conjunction
with a conventional rotary drilling technique in which the drill
string 18 is rotated at the surface, shock subs in general and the
shock sub of this invention in particular are quite desirable for
us~ with down hole powered drills in which the bit is turned by a
motor~suspended in the hole by a drill string which is not normally
rotated at the surface. Typical down hole motors of this type are
the turbine variety and the drill is known as a turbodrill.
Although the invention has been described in its preferred
form with a certain degree of particularity, it is understood that
the present disclosure of the preferred form is only by way of
example and that numerous changes in the details of construction and
the combination and arrangement of parts may be resorted to without
departing from the spirit and scope of the invention as hereinafter
claimed.
