Note: Descriptions are shown in the official language in which they were submitted.
78~9
Backgrourld of the Invention
a. Field of the Invention
This inverltion relat~s ~o earkh boring tools, and
more particularly to drill s~ring resilient units useful in
earth boring by the rotary system of drilling, such tools
sometimes being called vibration dampers or shock absorbers,
and relates speciflcally to a drill string ~plined resilient
t~ular telescopic joint.
b. Objects of the Invention
A principal object of the invention is to provide
such a tool especially adapted for balanced load drilling,
i.e. where the drilling weight and pump apart force are
equal. Though the tool is specially intended for balanced
load drilling, the tool may also be used with advantage
under such other load conditions as may normally be expected
and includes fea-tures o general utility in drill string
resilient units.
:
Other objects and advantages of the invention will
appear as the description thereof proceeds.
c. Description of certain Prior Art
It appears that many prior art drill string resil-
ient units have n~t been intended for use in balanced load
drilling. For example, in United States patent number
~57849
2,991,635 - Warren
there is disclosed a tubular telescopic tool employing
splines to transmit torque and multiple helical springs
disposed in parallel to transmit axial force. Three sets of
multiple sliding seals are employed, one above and one below
the spline-spring means, and one separatin~ the spline from
the springs. The uppermost multiple seal includes upper and
lower pairs of O-rings and a lubricating packing therebetween,
with a grease ~itt.ing to allow injection of grease. However,
the spring means come into action only upon contraction of
the damper from the no load condition. Warren states that:
"normally the springs 60 will be designed to
be compressed substantially half way during a
normal drilling operation. However, if a
greater or lesser weight is desired on the
drill bit 30, the drill string may be lowered
or raised a short distance to increase or
decrease the compression of the springs 60,"
(col. 5, lines 50-60).
:
In United States patent number
3,949,150 - Mason et al
there is disclosed a sealed, lubricated, splined, resilient,
tubular telesc:opic joint for a drill string wherein contrary
to the more usual arrangement the mandrel is connected to
the drill string and the barrel or case is connected to the
bit. A stack of rubber rings each sandwiched between flanged
_ 3~
57!3~9
metal rings serves as a resilient element. Referring to the
prior art it is said that due to hydrostatic pressure and
the difference in area between the upper and lower seals of
some dampers the resilient ele~en-t is preloaded in compres-
sion, necessitating the use of "hard" de~ormable elements
e.g. with an initial spring rate requiring 100,000 lb. for
the first 3/4 inch deflection and a later spring rate re~uir-
ing another 100,000 lb for the next 1/4 inch deflection,
but which will still go solid at normal drilling depths due
to hydrosta-tic pressure. Other dampers are said to avoid
this by using a floating seal for one end of the spring
chamber, so as to equalize the pressure, but in such dampers
Mason indicates that inappropriate springs were used.
Mason notes that for many shock absorbing elements
the spring rate increases with the load causing the damper
to operate at a point where the spring rate is high. Mason~s
objective is to provide a damper which operates at a position
where the spring rate is low at all times. Mason thereEore
discloses a shock absorber having a low initial spring rate
and having a fIoating seal to equalize pressure and eliminate
preload. But since his spring means is only single acting,
this requires operation about a mean position in which the
spring means is partially compressed, so that the spring
rate is higher.
Mason contemplates operating the damper with a
static spring force of about 25,000 lb., at which point the
damper has a spring rate of 20,000 lb. per inch. This
static load on the spring is said to be the difference
between 55,000 lb. drilling weight and a 30,000 lb. pump
~ 5~8'~
apart force. For operatiny in shallow wells where the bit
weight may be low, to insure some static compression of the
spring Mason proposes to reduce the pump apart force by
reducing the area of the seal between rnandrel and barrel.
To that end he provides a lower seal ring of smaller area
and vents the annulus between the lower seal ring and the
floating seal ring.
Mason further notes (at col. 12, lines 21 et se~)
that:
"As may occur in some cases, the area enclosed by
the outside diameter of the pxessure seal ring
times the differential pressure may be in excess
of the load to be carried on the bit. In this
case the damper would remain pumped open and
would not function as intended."
.~
To overcome this situation Mason eliminates the static
pressure balancing floating seal and under-fills the spring
chamber with lubricant so that the static pressure overcomes
the pump apart force and compresses the spring enough for it
to be operative.
In short, Mason et al disclose a single acting
damper intended to operate about a partially compressed
static load position thereby to avoid banging against the
travel limit stops upon imposition of dynamic loads.
The Drilco Industrial division of Smith Inter-
national, Inc. manufactures a vibration damper including
telescoping members with a shoulder on the mandrel in be-
~ S--
~S78~
tween two shoulders on the barrel formi.ng upper and lowerpockets in which are disposed shaped rubber sleeves to
provide variable spring rates. The upper sleeve is effec~
tive during normal operating conditions and the lower sleeve
is effective when the bik is liyhtly loaded and the pump
apart force exceeds the bit weight causing the damper to be
extended rather than contracted in the static condition.
The damper is intended especially for water well and other
shallow hole drilling wherein the pump apart force is small.
The lower sleeve is different from the upper sleeve, whose
spring rate increases much less rapidly with de~lection.
Only one or the other of the rubber sieeves is strained at
any one time. The spline-spring chamber is exposed to
drilling fluid. This construction is shown in a brochure
entitled:
bearing the notation "0976 Printed in USA", ''Drilco
Industrial SHOCK SUB (Vibration ~ampener)"
and in
U.S. patent number 4,139,994 issued ~eb. ZO, 1979
to George Abraitys Alther.
It is stated in the brochure:
" Vertical shock is absorbed by a large elaso-
meric element in compression. The element mater-
ial is specially compounded to provide optimum
characteristics of load carrying ability, fatigue
resistance, resilience and dampening. The element
is geometrically shaped to provide uni~orm "soft-
ness" at bit weights from near zero to 30,000 lbs.
A second, smaller elastomeric element is provided
_ 6 ~
-
to cushion reverse loading during severe rebounds
or when pulling from the hole."
~: :
::
~: :
~:~LS7~
S~mlmary of the Invelltioll:
AccordinK to the invelltion there is provided a drill string
vibration damper compris:in~ a splined tubular telescopic joint and
resilient means urging the jo;nt to a nelltral position upon both
extension and contraction of the joint with equal force exerted by
said resilient means upon equal departures from said neutral position
by extension and by contraction of the joint, said resilient means
having an increasillg spring modulus upon clepartures from said
neutral pOsitioll in both directions.
The resilient joint is provided with spring means which
has like characteristics upon both extension and contraction of
the joint. The spring means may comprise one or more pairs o
oppositely directed like single action spring means, but preer-
ably one or mora double acting compression springs are employed
wherein the same spring means is compressed whether the joint is
extended or contracted. In either case, the spring means will
have a variable spring rate, the spring rate being low over a
wide range o~ deflec-
~57~
tion o~ the damper in bo-th directions from the zero deflec-
tion condition and then increasing rapidly as the travel
limit of the joint,~approached, both upon extension and con-
traction of the damper.
For the spring means r stacks of springs, e.g.
elastomer-metal sandwiches or Belleville springs, (conical
washers), in series or series parallel disposition, are
preferred, since it is easier to achieve the desired var-
iable spring rate with such construction as compared, for
example, to a helical spin~. For support and protection the
spring rings are disposed in an annular chamber between the
telescoping members. To prevent binding, the rings are
guided by only line contact with one of -the telescoping
members and are out of contact with the o-ther member.
To reduce wear, the spring chamber, spline, and
telescopic guide bearings are sealed from the drilling fluid
and lubricated with oil.~ one end of the chamber is sealed
by pressure seal means capable of withstanding the pressure
differential between the interior and exterior of the joint.
The other end of the chamber is sealed by a floating seal
means which moves to a position in which there is no pres-
sure differential across it. This not only allows for
volume changes but leads to a balanced load on the spring
means since the fluid pressure acting on the damper is
therefore effective over the whole area within the pressure
seal means.
Magnetic and electrical means are provided to
check on the presence and condition of the lubricant and
therefore the soundness of the seal means. The seal means
t,~ _
~S713 ~9
both face downwardly -to avoid trappiny dirt; i.n other words,
the seal means are positioned so that -the direction o~ flow,
were any fluid to leak past the seal, would be upward.
Fvolution of the Invention
Applicant's concept:ion of a damper especially
adapted for balanced load drilling called broadly for a
variable spring rate double acting damper, that is, one
wherein low spring force resists bo~h contraction and exten-
sion of the damper, and higher spring force is effective as
the damper approaches its travel limits upon both contrac-
tion and extension of the damper.
Double Acting Dampers
Relative to double acting dampers ~roadly consi-
~:~ dered, the following United Stakes patents and publications
:~ are noted.
` 1,960,688 - Archer (1934)
2,325,132 - Haushalter et al
: 3,033,011 - Garrett
3,099,913 -~Garrett
2,727,368 - Morton (1955)
3,122,902 - Blair et al
3,254,508 - Garrett
:~ 3,323,326~- Vertson
3,447,340 - Garrett (1969)
3,503,224 - Davidescu
3,779,040 - Garrett
_ J~
~ 57~3 ~ 9
"Cougar Shock Tool" Cougar Tool Co Ltd. -
ante c. 1977.
In the foregoing constructions, though the joint is flexible
both upon extension and contraction, note needs to made of
the positions of the travel limit means, a greater travel in
contraction indicating a joint intended to operate ~ith the
spring means in compression in the neutral position~ For
further consideration, the above listed patents are divided
into four groups according to the nature o~ the spring
means, as follows:
a. Shear Type.
In the Archer patent there is disclosed a tele-
scopic drill string damper in which a rubber sleeve between
the telescoping members apparently is loaded in shear,
allowing the rubber to be stressed whether the damper is
extended or contracted. A similar construction is sho~n in
the Haushalter et al patent. Neither of these constructions
reguires a spline, tor~ue being transmitted through the
rubber.
Garrett patent number 3,033,011 shows a vibration
damper employing a telescopic joint and an elastomeric axlal
resilient element which also transmits torque, travel limit
stops, guides to take bending moment, and a sliding seal.
The damper is double acting in that both extension and
contraction place the resilient element in shear. A similar
construction adding an emergency clutch to transmit torque in
case of failure of the rubber sleeve is shown in Garrett's
patent number 3,Q99,918.
~57t3 ~1~
In the Davidescu patent there is shown a damper
employing sleeve type rubber metal sandwiches between the
mandrel and barrel. Tt appears that both extension and
contraction of the damper ~ill place the resilient elements
in shear.
In Garrett patent number 3,779,040 there is dis-
closed a vibration damper to be~ used between the drill steel
and power swivel of a boring machine, employing an elastomer
resilient unit.
All of -the foregoing shear type dampers would
appear to have fairly constant spring rates.
b. Alternate Tension-Compression Type.
Applicantls patent numbex 3,254,508 relates to a
bellows type vibration damper, the bellows serving to trans-
mit tor~ue, axial for¢e and fluid and being surrounded by a
sealed chamber in which is disposed a lubricant separated
from the drilling fluid by ~he bellows and a floating seal.
The damper bellows is compressed upon contraction of the
joint and placed in tension when the damper is extended.
In applicant's patent number 3,447,340 there is
disclosed a damper similar to that of the aforementioned
Garrett patent number 3,254,508 except that the bellows has
helical convolutions. The travel limit stops are positioned
so that the damper can both extend and contract, tensioning
or compressing the spring, but the permissible extension
travel is much less than the contraction allowed.
~ ~ 2 _
~S78 ~
Although a single spring rneans is ernployed in the
double acting dampe.rs o the above two patents, the spring
action is different upon extension than it is upon contrac-
tion of the damper, and in bot~l cases the initial spring rate
at zero deflec-tion is apparently of the same order of magni-
tude as at the travel limits.
c. Diferently Positioned Springs Alternately
Compressed.
In the Morton patent there is disclosed a vibra~
tion damper to be used in propeller shafts. The damper
comprises telescoping members with a rubber sleeve in the
annulus therebetween and bonded thereto to t.ransmit torque.
Shoulders on the members engage the rubber sleeve axially so
as to transmit thrust. The arrangement appears to be such
as to axially compress one part of the rubber sleeve upon
thrust in one direction corresponding to contraction of the
damper and to axially compress another part of the rubber
sleeve upon extension of the damper (e.g. during reversal of
the propeller), and it appears that tension is alternately
applied to such parts of the rubber sleeve as are not com-
pressed. It is further stated that due to the differing
axial thicknesses of the sleeve there will be two natural
periods of torsional vibration so that torsional vibrations
will tend to interfere and cancel out.
In the Vertson patent there is disclosed a tele-
scopic damper including helical shoulders on the barrel and
mandrel with two joined helical rubber elements therebetween.
Apparently one element is compressed wl~ile the other is
~S7~3 ~9
unloaded, causing rubber to flow from one element to the
other, rega~-dless of whether the damper is extended or
contracted.
The Cougar leaflet appears -to disclose a tele-
scopic tool employing two rubber sleeves as resilient ele-
ments, which may act alternately upon extension and contrac
tion of the joint or may merely function as in United States
patent number
3,660,990 - Zerb
wherein it is said that a rubber seal in series with the
main resilient sleeve also acts to damp shocks.
It does not appear that any of the rubber elements
of the constructions of the foregoing group of patents is
designed to have a variable spring rate with an extremely
low spring rate over a range of deflection in both direc-
tions coupled with a very high spring rate at both travel
limits.
Variable Spring Rate Dampers
Applicant's first conception of a variable spring
rate resilient unit for balanced load drilling contemplated
the possible use of felted steel wire ("steelwool") annular
pads for the spring elements, such elements filling the
annular chamber between mandrel and case and engaging the
side walls thereof. Somewhat similar elements are shown in
United States patents number
~;78~9
3,383,126 - Salvatori et al
3,406,537 - Falkner, Jr.
Salvatori et al shows a vibration damper employing an upper
slidiny lip seal and a lower floating seal, with a spline
and a resilient element in the chamber therebetween. A
similar construction is shown in the Falkner, Jr. patent.
Falkner, Jr. fills his chamber with oil.
Double Acting Dampers With Single Spring
When appllcant first conceived of his invention of
a variable spring rate double acting damper for balanced
load drilling, he requested one of the engineers in his
department to draw up an embodiment. The engineer chose as
spring means for the double acting damper, a double acting
compression spring construction, thereby to compress the
same stack of wire pads upon both extension and contraction
of the damper. The wire pads were disposed between shoulders
on the barrel and mandrel, with spacers allowing the shoulders
to lie in different planes. Such a double acting spring
construction is employed in the preferred embodiment of the
double acting damper shown herein. Relevant to such con-
struction are the followin~ p~lications:
U.S. patent number 3,381,780 - Stachowiak (1968
USSR inventor's certificate number 1,279,335 -
Shprenk ~1970),
~57~
both of which apparently show such shoulder disposition, and
Shprenk appearing to show such spacers.
The Stachowiak patent relates to a formation
testing shock absorber. The shoclc absorbing element includes
an elastomer sleeve axially compressible and radially expan-
sible to pump a liquid out of dash-pot space. A solid
floating piston for pressure and volume compensation lS
employed. The damper is double acting, the mandrel and
barrel members each having two axially spaced apart shoul-
ders which extend less than half way across the annulus
between the members, so that the rubber sleeve captured in
the annulus between the shoulders is compressed axially upon
both extension and contraction of the joint. The resil~
iency of the sleeve is said to return the tool to full
extended position. When the elastomer sleeve is compressed,
it rubs on the surrounding case. Though the sleeve is
shaped, the efect is said t~o cause increasing friction as
deflection increases, without mention of any spring action.
At the same time, the deformation o~ the sleeve constricts
the dash pot action. A constant force versus deflection is
said to be achieved. This corresponds t~ a constant spring
rate of zero.
The~Shprenk et al publcation, entitled "Superbit
Shock Absorber", appears to disclose a vibration damper
employing barrel and mandrel members with elastic rings
gripped in parallel between the members and engageable by
axially spaced shoulder means on the mandrel and also by
axially spaced shoulder means on the barrel, to strain the
elastic elements whether the damper is extended or contracted,
~57l3 ~9
spacers allowing the shoulders to be non-coplanar. However,
the elastlc elements are strained, apparently, by relative
axial rnovement of their inner and outer peripheries relative
to a bridging washer embedded in the elastic elements rather
than by axial approach of their end faces.
.~
In neither the Stac~owiak nor the Shprenk publica-
tions is there any discussion of the pump apart force nor
balanced load drllling nor any disclosure of using a plural-
ity of rings in series to achieve high flexibility, nor
paralleled groups of series stacked rings to reduce stress on
individual rings.
Rubber Sandwich Springs
In the evolution of his invention, applicant later
conceived of a spring means including a plurallty of elas-
tomer rings of oval cross-section sandwiched between dished
metal washers. A search relative to this type of spring
revealed certain prior United States patents as follows:
2,733,915 - Dentler (1956)
2,930,491 - cOO]; (19SO)
2,946,462 - Danielson (1960)
2,982,536 - Kordes (1961)
3,330,519 - Thorn ~1967)
3,480,268 - Fishbaugh (1969)
3,493,221 - Mozdzanowski (1970)
3,677,535 - Beck (1972)
3,814,411 - Aarons et al (1974)
3,997,151 - Leingang (1976)
~.~ 571 3 ~9
Applicant has since conceived of an improved spring rneans of
this type providing for paralleling several groups of such
rubber sandwiches, as will be described ln more detail
hereinafter.
Other drill string resilient units employing
rubber sandwich spring means are known. For example, there
is disclosed in British patent number
Br. 220~70 (1924) - Reichwald
a telescopic jar with torque transmissin means, useful for
well drilling, wherein the jar is cushioned on the lifting
or jar extension stroke by a plurality of rubber rings
sandwiched between flanged metal rings. The flanges, nor-
mally out of contact, limit compression of the rubber. The
rubber rings are spaced radially from both the inner and
outer telescopic members and the flanges of the metal rings
surround the inner member~
:
A tubular telescopic joint employing a spline,
sliding seal, and resilient means, to form a vibration
damper is shown by United States patent number
3,301,009 - Coulter, Jr. (1967)
In the Coulter construction the resilient means includes a
plurality of hexagon cross-section elastomer rings with
single flat washers interposed between each ad~acent pair of
elastomer rings. The resilient means is single acting. The
sides of the rings rub against the mandrel of the telescopic
~ 5~ g
joint. ~ form of inver~ed floating seal is shown at the
upper end of the spring cham~er, but -the spriny char~er is
not sealed at its lower end.
Belleville Springs
In the further evolution of his invention of a
resilient unit for balanced load drilling, applicant con-
sidered the employment o~ Belleville springs disposed in
series-parallel stacks. At this stage, applicant was par-
ticularly concerned with protecting the seals employed for
the lubricated chamber in which the springs are moun-ted.
The use of plural telescopic joints, one for each
leg of a three cone roller bit is shown in United States
patent number
~; ,
2,815,928 - ~odine,
the joints each including a vibration damper in the form
(Figure 6) of a stack of dished washers alternately dis- ~
posed. Axially spaced cylindrical bushings provide vertical
motion supports. Interengaging means on the bit legs and
bit body limi-t relative rotation. In a modification (Figure
9), a single spline telescopic joint is used and the resil-
ient member is in the form o~ a long steel sleeve. The
spline is in a sealed chamber filled with grease~ the seals
for the chamber are sliding seals. The steel sleeve is also
sealed at both ends and extends above and below the spline.
Volume compensating ports at one end expose the upper spline
seal to well fluid and the lower spline seal is exposed
thereto at the lower end of the tube.
_ / q _
~57~3~g
In Uni.ted States patent number
3,539,026 - Sutlif~ et al
is disclosed a fishing tool energizer comprising a splined
telescopic joint with yroups of paralleled Belleville springs
stacked in series to provide the resilient unit. To vary
the spring force, more or fewer springs can be parallel in
each group.
A vibration damper employing ~elleville springs
for the resilient element is shown in United States patent
number
3,871,193 - Young.
The resilient element and spline are in a lubricating oil
chamber formed by upper and lower sliding seals between
telescoping inner and outer members. The resilient element
includes a plurality of spring means of diferent spring
rates stacked in series. When the lighter spring means
reach their limit of travel, the heavier spring méans are
left, presenting a bigher spring rate. In two embodiments,
groups of Belleville springs are employed stacked in series
with varying members of springs paralleled in each group.
In another embodiment coil springs of differing cross-sections
are stacked in series.
United States patent number
_ z~._
~ L57~3~9
3,898,815 - Young
discloses a sealed, lubricated telescopic tool including
ball spline means and disc spring means.
A vibration damper made in Canada by Smith Inter-
national Canada, includes a telescopic joint, spline, a
stack of several groups of Belleville springs with the
groups in series and khe springs in parallel in each group,
seals above and below the spline providing a cha~ber for
lubricant, a three inch travel from no load to 60,000 lb.
and a spring rate of 20,000 lb/inch.
None of the above constructions appears to be
concerned with balanced load drilling.
Most recently, applicant asked J. B. Hiebel, one
of the men in his department, to further the development of
his invention of a double acting damper for balanced load
drilling. This further development resulted in a damper
construction employing roller Belleville springs, shown as
the preferred embodiment of applicant's damper for balanced~
load drilling. A separate application by J. B. Hiebel on a
:
damper employing roller Belleville springs is contemplated.
Roller Belleville springs are a species of vari-
able moment Belleville springs, the latter being disclosed
in United States patent number
2,675,225 - Migny.
IL~;57~
In Figure 3, it is sho~m that .such spri.ngs rnay be stacked in
series-parallel, but the parallel springs are not all of the
variable moment t~pe, including simple dished washers in the
middle of each parallel group.
SEALED l.UBRICATED SPRING~-SPLINE-BEARING CHAM~ER
As noted previously, part o applicant's overall
concept of a damper for balanced load drilling included a
sealed and lubricated spring~spline-bearing chamber. In
addition to the patents previously discussed, a number of
other patents may be mentioned relative to the employment of
a sealed lubricated spring-spline-bearing chamber, in a
tubular, axially resilient telescopic joint for well drill-
ing, as follows:
2,025,100 - Gill et al
2,240,519 - Reed (three lobed spline)
3,323,327 - Leathers et al
3,345,832 - Bottoms
3,581,834 - Kellner and Garrett
7~ C~
Brief Description of the Drawinys:
For a description of a preferred er~odiment of the
invention, reference will now be made to the accompanying
scale drawings, the elevational and cross-sectional views
showing that the paîts are made of metal, e.g. steel, except
as otherwise indicated, e.g. the seals are prefexably made
,oo~ f 'e ~4 7/~ /e,Je
o~ ~e~ or an oil and water resistant elastomeric mater-
ial, and in a modifi.cation elastomer-metal sandwich spring
elements are employed.
FIGURE 1 is an elevation, partly in section,
showing a damper according to the invention assembled in the
lower end of a drill string, with the spring means in the
neutral or unloaded position.
FIGURE 1~ is a detail to a larger scale, showing
the electrical test probe o~ the damper;
: .
FIGURES 2A and 2B, together hereinafter referred
to as FIGURE 2', are fragmentary sectional views to a larger
scale showing the pressure seal, bearing and spline at the
lower end of the damper, with the damper in the open or
extended position;
FIGURES 2A" and 2B" together hereinafter referred
to as FIGURE 2", are fragmentary sectional views similar to
FIGURE 2' showing the lower end of the damper in the closed
: or contracted position;
_ 2 3 -
~57~3 ~C~
FIGURES 2' and 2" are drawn with a common center
line for easy comparison and may be referred to toyether as
FIGURE 2;
FIGURES 3A', 3B' and 3C', together hereinafte~
referred to as FIGURE 3', are fragmentary sectional views,
also to a larger scale than FIGURE 1, showing the resilient
unit at the medial portion of the damper in its open or
extended position.
FIGURES 3A/', 3B" and 3C", together hereinafter
referred to as FIGURE 3", are fragmentary sectional views,
similar to FIGURE 3', showing the medial portion of the
damper in the closed or contracted position.
FIGURES 3' and 3" are drawn with a common center
line for easy comparison, and these flgures taken together
may be hereinafter referred to as FIGURE 3;
: .
FIGURES 4A' and 4B', together hereinafter referred
to as FIGURE 4', are fragmentary sectional views, also to a
larger scale than FIGURE 1, showing the pressure volume
compensating ~loating seal at the upper~end of the damper
:~ being in the open or extended position;
FIGURES 4A" and 4B", together hereinafter re~erred
to as FIGURE 4", are views similar to FIGU~ES 4A' and 4B',
showing the upper:end of the damper in the closed or con-
tracted position;
~L~57~
FIGURES 4' and 4" are drawn with a cor~on center
line for easy comparison and may be referred to together as
FIGURE 4;
FIGURES 5, 5A and 6 are respectively sections
taken at planes 5-5, SA-SA and 6-6 of FIGURES 3 and 2';
FIGURE 7 is a sectional view of one of the roller
Belleville spring elements;
FIGURE 8, 9 and 9A are graphs; and
J FIGURE 10 is a fragmentary sectional view showing
' a modified form of resilient means.
.
_æ ~ ~
1~78~
Descript:ion of Preferred Embodiments:
- Damper -
Referrlng now to the drawings and in particular to
FIGURE 1, there is shown a damper 21 connected at its lower
end to a three cone drill bit 23 and at its upper end to
another lower drill string member 25 such as a skabilizer or
drill collar. Although the drawing shows the damper dir-
ectly above the bit, it may be employed at o-ther places in
the drill string, preferably however where most of the
weight slacked off in the drill string is above the damper.
The apparatus is shown in a well bore 27 being drilled by
bit 23 by the rotary system of drilling. The damper in-
cludes a tubular mandrel identified generally by reference
number 29. The mandrel comprises a lower seal and bearing
and spline portion 31, a medial spring portion 33, a cross
over sub portion 35, and an upper compensator portion 37.
Mandrel portions 31, 33, and 35 are connected by rotary
shouldered taper threaded connections, such connections
- being more fully described in United States patent number
3,754,609 ~ Garrett.
:~ :
Mandrel portion 37 is shrink fitted to mandrel portion 35.
Mandrel 29 works telescopically within and about a
tubular barrel indicated generally by reference character;
39. The barrel includes a lower seal and bearing and spline
portion 41, a medial spring portion 43, an upper cross over
sub 45, and a depending compensator portion ~7. Barrel
portions 41, 43, ~5 are connected together by rotary shoul-
~L~L57~
dered connections as above referenced. sarrel portions 45,47 are connected together by a taper thread and are sealed
by an O ring 49.
The connections between the damper and drill bit
and between damper and the uppex part of the drill s-tring
are also r~tary shouldered connections. Such connections
each comprise a pin and box co~lector and either type of
connector may be provided at each end of the damper, depend-
ing on what type of use is to be made of -the damper. As
shown, a box connector ~ is provided at the lower end of
the damper for connection with a drill bit pin ~, and at
the upper end of the damper there is a box connector 44 for
connection with a pin 46 on lower drill string member 25.
Passage means through the damper for conveylng
drilling fluid ~rom lower drill string member 25 to drill
~it 23 include central passages 48, 50 through tubular
mandrel portions 31, 33.
Referring now to FIGU~E 2, there is shown most of
the lower portion of the damper, forming the seal, bearing
and spline part thereof. This includes portions 31, 41 of
the mandrel and barrel.
- Pressure Seal -
At the lowermost part of the damper there is seal
means 51 disposed between the barrel and mandrel portions
31, 44. Seal means 51 comprises a plurality of double lip
seal rings 53, 55, 57, 59. Seal rings 53~ 55 face upwardly;
seal rings 57, 59 face downwardly. The seal rings are
~1578'~J
c 7~-q ~ O ~o e f ~
preferably made O~ G~ or other low friction coefficient,
high temperature resistant, flexible, resilient, sealing
material. The lips of the seal rings are preloaded to move
away from each other by corrosion resistant metal springs
such as those indicated at 52, 54, 56, 58.
Metal wedge rings 61, 63, 65, 67 also hold apart
the lips of the seal rings to assist them in moving into
sealing engagement with low friction coefficient hard ~etal
finished surface 68 at the cylindrical outer periphery of
mandrel portion 31 and the inished surfaces 70, 72, 74 at
the cylindrical inner periphery of barrel portion 41 and the
cylindrical upper and lower inner peripheral portions of
spacer ring 76. Weep holes 60, 63, 64, 66 equalize fluid
pressure on opposite sides of the rings. The wedge rings
have tongues extending to the bottoms of the annular spaces
between the seal ring lips to transmit force -through the
bottoms of the seal rings when axial force is imposed on the
wedge rings. Although as noted above, -the wedge rings
spread apart the lips of the seal rings, their function of
transmitting force through the bottoms of the rings being
their primary function, thus permitting stacking of the seal
rings, as in the case of rings 53, 55. In addition, the
flat surface of the uppermost wedge xing 61 facilitates
retention of the seal rings in the barrel underneath steel
retainer ring 69. Ring 69 is beveled and rests against
bevel shoulder 71 in the barrel.
The force of the fluid pressure in the damper
acting down on seal rings 53, 55 is transferred by end ring
73 through support ring 76 to end ring 77 and thence
~571~9
to snap ring 79, received in annular groove 81 in the lower
end of the barrel.
Support ring 76 is sealed to the barrel by O rings
82, 83 received in annular groove around the support ring.
The o rings are in slight compression between the boktoms of
the grooves and barrel surface 8$. support ring 76 is held
fixed in the barrel be-tween end rings 73, 77, wh~ch are
captured between snap ring 79 and a shoulder 87 at the
juncture of upper seal surface 70 and larger diameter lower
seal surface 85. Therefore O rings are suitable for the non-
sliding seal between the barrel and support ring 76. This
is in contrast to the seals effected by seal xings 53, 55,
57, 59 with the mandrel, which are axially sliding seals.
Support ring 76 not only transfers load from the
upper seaI rings to snap ring 81 ~ut also provides a car-
tridge independently supporting seal rings 53, 55 so that
load is not transferred from one through the other as in the
case of the uppermost seal rings 53, 55, thereby insuring
that the lip seal action of each ring remains unimpaired. A
similar cartridge construction could be used for the upper
two seal rings if desired. Or if preerred, the lower seAl
rings 57, 59 could be stacked with only a wedge ring in
between as in the case of the upper seal rings.
The lower seal rings 57, 59 seal primarlly against
upwardly directed fluid pressure from the fluid outside the
damper. The force of the pressure is transferred to shoulder
87 through end ring 73 in the case of seal 57 and through
support ring 76 and end ring 73 in the case of seal ring 59.
_ ~q ~,
.
~S~ 9
No force of pressure fl-lid is intended to be transferred
from end ring 77 to wedge ring 66, nox from support ring 76
to wedye ring 65.
- Test Probe -
As will appear more clearly hereinafter, although
seal means 51 seals against the pressure of fluid both in
the well bore outside the damper and in the drill string
connected to the damper, seal means 51 is exposed to dxill-
ing fluid only on its lower side. Above seal means 51 in
the space between the barrel and mandrel there is a clean
lubricating oil 91 extending all the way up to the compen-
sator portions of the barrel and mandrel. Within this clean
fluid work the spline and resilien-t means later to be de-
scribed. It is important for the user to know if any of the
damper seals has failed. If such ~ailure causes an influx
of drilling fluid, the lubricating fluid will become con-
taminated by the drilling fluid, e.g. with water.
Referring now particularly to Figures 1, lA, and
2A', to detect such a change there lS provided in the barrel
just above lower seal means 51 a test probe comprising an
electrically conductive (metal e.g. brass) electrode 93
extending through the barrel wall. The electrode is sur-
: `rounded on its outer periphery by and bonded and sealed to
electrically insulating sleeve 97 which in turn is mounted
in and bonded and sealed to screw plug 99 which is the
closure for a lubricant ~lnjection and drainage port provided
by threaded hole 101 in the barrel wall. The plug is screwed
into the port. To prevent the electrode from moving axially
under the pressure differential between the interior and
- 30
.
~S7~
exterior of the barrel, the electrode is made of larger
diameter at its mid por-tion than at its ends, leaving
tapered shoulders 98, 100 ad~acent each end. An annl1lar
recess in the screw plug is shaped correlative to the
exterior of the electrode forming shoulders at each and of
the recess facing toward the shoulders on the electrode.
The insulation sleeve is captured at each end between the
plug and electrode shoulders and resists relative motion of
the electrode and plug. The outer diameter of the mid
portion of the electrode is too big to pass through the
recess in the plug at the outer end thereof. A ma-terial
suitable for the insulation sleeve is one which can with-
stand pressure differential and well fluids such as a
plastics material, e.g. epoxy. At its outer end, the recess
in plug 99 is formed as a wrench socket 102 to facili-tate
assembly and disassembly.
By connecting an ohmmeter 103 between the outer
end of the electrode and a point such as 105 on the exterior
of the barrel, the electrical resistance of the oil 91 can
be measured to determine its character, i.e. whether or not
it has become contaminated. If so, the damper seals and
lubricant need to be checked for replacement and then re-
placed to the extent required.
It may be noted that whereas the lubricating oil
has a high resistivity, most drilling fluids have a low
resisti~ity. Furthermoxe, the drilling fluid will usually
be denser than the oil and will sink to the bottom of the
space normally occupied by the oil. That is why the test
probe is placed at the lowermost part of such space, just
3 1 ~
~L~S78~9
above lower seal means 51. At that point, the test probe
will coIltact any such dr.illing fluid and -the resista~ce test
will show a low resistance path through such drilling fluid.
- Lower Bearing Means -
Still referring to FIGURE 2, above plug 99, the
interior of barrel portion 41 is provided with a replaceable
bushing or liner 121 made of a wear resisting, low friction
coefficient, corrosion resistant bearing material compatible
with hard facing m~terial 68 on mandrel 31. For example,
liner 121 may be made of beryllium copper or aluminum bronze
or the equivalent. The bushing has a smooth cylindrical
inner surface which cooperates with a continuation of sur-
face 68 on the exterior of the mandrel to provide bearing
means. The bearing means transmits bending moment betwe2n
mandrel and barrel while providing for relative axial motion
therebetween. To provide maximum area for taking bending
moment, flutes 147 are made deeper ~han they are wide.
: ~:
.
- Spline -
Above the bearing means just described, the wall
of barrel portion 41 thickens by a reduction in its inner
diameter, and there is a correlative thinning of the wall of
mandrel portion 31 by a reduction in its outer diameter.
Conical surfaces 123, 125 are formed where these transitions
occur.
: ~
~78'~g
Referring now al~o to EIGURE 6, the interior of
the thic~ walled part of barrel portion 41 is fluted ~aral-
lel to the barrel axis, forming three vertical grooves 131,
133, 135. The mandrel at this level is provided with three
splines 137, 139, 141 which it into the grooves and cooper-
ate therewith to provide spline means fox transrnitting
torgue between the barrel and mandrel while providing for
relative axial motion. While a spline can be made to trans~
mit bending moment, e.g. by having the spline bottom in the
grooves, the spline means here disclosed is intended primar~
ily for transmission of tor~le through the side walls of the
splines and grooves, the walls having large radial compon-
ents. However some transmission of bending moment will also
be provided by the spline means due to the circumferential
components of the side walls of the splines and grooves.
Each spline side wall may, for example, make an angle of
e.g. 15 to 45 with the radlal plane therethrough, Figure 6
sho~ing a representative desirable angle.
:
- Upper Bearing Means -
Above the spline, just described, the inner peri-
phery 143 of buried pin 38 is in sliding contact with the
; outer periphery 145 of mandrel portion 31, thus providing an
upper guide bearing. The upper and lower guide bearings,
being spaced apart axially, can take considerable bending
moment without being overloaded. The outer periphery of
mandrel 31 is fluted at 147, providing fluid passages for
lubricating oil 91, as will be explained in more detail
hereinafter.
_ 3~3
~.~57~3~9
- Resil-en~ Means -
Referring now to FIGURE 3, -there is shown the
medial portion of the dampèr including the resilient means
thereof comprising mandrel port:ion 33 and barrel portions
42, 43 with spring means 151, 153 therebetween. M~ndrel
portion 33 has an outturned flange 155 (Figure 3B1, 3C')
therearound, and buried pin 40 between barrel portions 42,
43 forms in effect an inturned flange within -the barrel~
These flanges provide upwardly facing annular shoulders 158,
159 and downwardly faciny annular shoulders 160, 161. These
shoulders define the inner ends of annular pockets 162, 16
between mandrel portion 33 and barrel portions 42, 43. At
the inner ends of these pockets are pressure rings 164, 165
engaged respectively with the ends of spring stacks 167, 169
and engageable ~ith shoulders 158-161.
The upper end of pocket 163 i5 defined by annular
shoulder 171 on barrel portion 43 and annular shoulder 173
provided by the lower end of mandrel portion 35. Pressure
ring 174 is engaged with the upper end of spring stack 169
and is engageable with shoulders 171, 173~
The lower end o:f pocket 162 is defined by annular
shoulder 175 provided by:the upper end of buried pin 38
connecting barrel portions 41, ~2, and annular shoulder 177
formed by the upper end of mandrel portion 31. Pressure
ring 183 is engaged with the lower end of spring stack 167
and is engageable with shoulders 175 and 177.
- 3~f_
~S7~3 ~9
It will be apparent that for each spring means
151, 153 the mandrel is provided ~Jith upper and lower shoul-
ders and the barrel is provided with upper and lower shoul-
ders, and that since the mandrel and barrel shoulders do not
overlap, they can pass each other whichever ~ay the barrel
and mandrel are moved axially relative to each other.
Regardless of whether the dampe:r is contracted or extended,
pairs of mandrel and barrel shoulders will engage the pres-
sure rings to cause the spring means to be compressed axial-
ly. Such action is indica-ted in the left and right hand
halves of Figure 3.
- Neutral Position -
Referring now to Figure 1, the damper is shown in
the unextended, uncontracted, or neutral position in which
the spring means are of maximum length. In the neutral
position, the shoulders 158,- 160 on the mandrel flange 155
tFig. 3B', 3C') are co-level or in alignment with shoulders
159, 161 on barrel pin 46. At the upper end of spring means
153 (referring to Fig. 3C' and 3C" for reference numbers but
not for position), in the neutral position pressure ring 174
is engaged both with mandrel shoulder 173 and barrel shoulder
171. Similarly (referring to Figures 3At and 3A" for refer-
ence numbers but not for position), in the neutral position
pressure ring 183 is engaged both with mandrel shoulder 177
and barrel shoulder 175.
-
- ~5
~57~i~9
- Travel Limits ~
A nut 189 i5 screwed onto a straight thread on the
exterior of the upper end of mandrel portion 31. Nut 189
provides shoulder 191 at i-ts lower end to engage barrel
shoulder 175, forming therewith stop means limiting exten-
sion of the damper. Nut 189 is provided wi-th a collar 193
which makes an interference (shrink) fit with the uppermost
part of mandrel portion 31. A thin section lg5 between
collar 193 and the body of nut 189 facilitates sawing off
the collar if the nut needs to be removed.
Pressure ring 183 is provided with an engageable
surface 207 to enyage shoulder 177 and an annular tongue 209
whose lower end 210 is adapted to engage shoulder 175.
Tongue 209 spans nut 189 and the threaded connection between
mandrel portions 31 and 33. Although the latter connection
is a rotary shouldered connection as previously mentioned,
it is also provided with an O ring seal 211 in case there is
difficulty in applying ade~uate make up torque to the con-
nection to effect a seal at shoulders 213, 21~. Screw
threads 217, 219 are straight (untapered) threads to facili
tate make up.
The L-shaped cross-section of pressure ring 183,
which provides tongue 209, permits shoulders 175 and 177 to
be in different planes, i.e. non-coplanar, as occurs, e.g.
because of the presence of nut 189.
Means limiting contraction of the damper is pro-
vided by stop shoulder 208 ¢Figure 2A') on the lower end of
-~G ~
7 ~
the barrel and stop shoulder 210 on the lower part of the
mandrel. It will be seen that the possible contraction of
the darnper from the neutral position equals the possible
extension from the neutral position. However they could be
made unegual if deslred.
-Pressure-Volume Compensation Means-
Referring now in part to FIGURE 3C' and more
particularly to FIG~RE 4, there is shown the upper part of
the damper including mandrel portions 35 and 37 and barrel
portions 43, 45, and 47. An annular volume 231 is formed
between mandrel portion 37 and barrel portion 47. Due to
the presence of cross ovèr mandrel portion 35, mandrel
portion 37 is of larger inner diameter than the cuter dia-
: meter of the wash pipe formed by barrel portion ~7; in other
words, mandrel portion 37 is outside of barrel portion 47.
Barrel portion 47 forms a part of and a continuation of the
drilling fluid conduit means through the damper, which
means, as previously mentioned, includes the passages 48, 50
through the interiors of tubular mandrel portions 31, 33.
Annular volume 231 between barrel portion 47 andmandrel portion 37 is closed by floating seal means 233
comprising tu~ular metal cartridge 235 carrying a plurality
of sliding seal elements. Because barrel portion 37 is
outside mandrel portion 47, an inversion of the usual barrel
and mandrel relationship, any drilllng fluid tending to flow
through volume 231 must flow upwardly. Therefore detritus,
sand and other particula~es carried by the drilling fluid,
when stopped by seal means 233, will fall down out of vol~me
- 3 7 -
~57t3~9
231) away from seal means 233.
Summarizi.ngJ the:re is no problem o~' drilling ~luid
particulates accumulQti.ng above the seal means at either the
lower or upper end of the damper, since the spaces above
these seal means arc filled with lubricating oil and since
a~ the lower faces of these seal means any drilling fluid
particulates will fall out of the seal means due to the
orce of gravity.
Floating seal means 233 includes double lip,
spring loaded seal rings 237-240 on the interior of car-
tridge 235 to seal between the cartridge and the outside o
wash pipe barrel portion 47. Similar seal rings 241-244
seal between the outside of the cartridge and the interior
of mandrel portion 37. Seal rings 237-244, similar to the
seal rings in seal means 51 at the lower end of the damper,
are provided with preload springs 245-252 to press their
lips apart into sealing engagement with the cartridge and
barrel and mandrel portions. Also, wedge rings 253-260 are
provided to allow for stacking the seal rings in series, to
transmit forces from one seal ring to another, and to facil-
itate retention. The wedge rings are provided with weep
holes, e.g. as shown at 261, for pressure balancing as in
the case of the wedge rings forming parts of the sealing
; means at the lower end of the damper. The seal rings and
~: :
~; - 38 -
3L~lS78 ~9
wedge rings are retained on the cartridge against back up
flanges 265, 267 by end rinys 269~272 and snap rings
274-277, the latter being received in annular grooves in the
ends of cartridge 235.
The whole seal means 233 is free to move as a unit
axially up and down within volume 231, travel being limited
by the upper end of the cartridge engaging annular shoulder
278 on washpipe barrel portion 47 and by the lower end of
the cartridge engaging shoulder 279 on the upper end of
mandrel portion 35. In normal operation the cartridge will
never engage the stops. As long as the seal means is free
to move, there is no pressure differential across it. It
moves up or down so as always ~o be in contact with and sup-
ported by lubricating fluid 91 that fills the lubricant
space between the barrel and mandrel in between pressure
seal means 51 and floating seal means 233. Floating seal
means 233 thus provides a pressure-volume compensator
accomodating to changes in the volume of the lubricant
space, allowing lubricating oil 91 to flow into the space
283 between washpipe barrel portion 47 and mandrel portion
37 when the lower part of the lubricant space between the
barrel and mandrel reduces, and causing lubricating oil 91
to flow back into the space 283 when it enlarges, while
keeping the lubricating oil separate from the drilling fluid
at all times.
-Lubricant Space - -
The space occupied by lubricating oil 91, extend-
ing from the lower part of the damper to the upper part
~ ~ 9 _
~ S7~ ~9
thereof, may be traced from just above pressure seal means
51, past test probe electrode 93, between liner 121 and
bearing surface 68, into space 281 between conical portions
123, 125, in between splines 131, 133, 135 and grooves 137,
139, 141, thr~ugh flutes 147, throuyh channels 282, 284 cut
across the lower ends of nut 189 and tongue 209 (e.g. when
the nut and tongue engage shoulcler 175 as in Figure 3A'),
p.ast the outer and inner peripheries o~ ring 183 (and its
tongue 209) inside barrel porkion 42 and outside mandrel
portion 33, through spring pocket 162 inside and outside
spring stack 167, between pressure ring 165 and shoulder 160
or 161 (one or the other is out of contact with the ring
except in neutral position), or in neutral position through
radial channels 286 ~or 288) across the uppermost face of
ring 165 as it happens to be assembled, through annular
space 156 between barrel and mandrel, between pressure ring
164 and shoulders lS9, 158 ~one or the other is out of
contact with the ring except in neutral position), or in
neutral position through radial channels 287 (or 289) across
the lowermost face of ring 164 as it happens to be assem-
~led, through spring pocket 163 inside and outside spring
stack 169, between ring 174 and shoulder 173 when out of
contact or between ring 174 and shoulder 171 when out of
contact (one or the other shoulder is out of contact except
in the neutral position), or in neutral position through
radial channels 283 (or 285) across the uppermost face of
ring 174 as it happens to be assembled, through annular
spaces 204, 206, 208, and into uppermost space 283. : -
_ 1~ 0 - .
L 578L~9
- Motion of Floating Seal -
It is to be noted that when the damper is in use,
the desired end result is zero axial motion of the barrel,
which is connected to ~he upper part of the drill string,
despite up and down motion of the mandrel, which is con~
nected to the drill bit. As the mandrel moves up and down
the volume of the space between the parts of the mandrel and
barrel delimited by the lower seal rneans and the volume
compensator changes and the compensator moves up and down to
accomodate for the volume change. Large volume changes .
occur at space 281 between conical surfaces 123, 12~5 ~FIG.
2) and at space 283 ahove the upper end of mandrel portion
37. Floating seal means 233 (Figure 4) therefore moves up
and down rapidly relative to both barrel portion 47 and
mandrel portion 37 during operation of the damper. In the
embodiment shown, the axial travel of the floating seal is
about 3/5 of the axial travel of the mandrel relative to the
barrel. For this reason the outer periphery of washpipe
barrel portion 47 and the inner periphery of mandrel portion
37 are each provided with a hard metal coating, e.g. nickel
plated, as shown at 275 and 276, such coating having a low
friction coefficient and a smooth finish. Both wash pipe
barrel portion 47 and mandrel portion 37 are readily replace-
able should they become unservicable for any reason.
Although floating seal means 233 moves rapidly up
and down to accomodate changes in volume of the space occu-
pied by the lubricating oil, it may also move up and down
more slowly in response to changes in the volume of the oil
itself as temperature and pressure change.
.
57~
- Seal Position Indicator -
Still reerring to FIG~RE 4, on the upper end of
cartridge 235 is disposed a cap 291 having a skirt 293. A
plurality of screws 296 circumferentially spaced apart
around the skirt extend through holes in the skirt into
threaded holes in the cartridge. The upper end 297 o~ cap
291 has a conical top face pointing upwardly. Surmounting
the cap is permanent magnet ring 299, which has a lower
conical face correlative to the top face of cap 291, being
suitably secured thereto, e.g. by epoxy cement or by sinter-
ing or solder.ing. ~agnet ring 299 is magnetized radially,
whereby its inner periphery is of one polarity and!its outer
periphery is of an opposite polarity. Cartridge 235 and
wash pipe 47 are made of non-ferro-ma~netic ma-terial, such
as stainless steel. A magnetic probe, such as a steel
building stud locator, lowered into wash pipe 47, will indi-
cate the level of the magnet ring and hence of the floating
seal. If the damper is fully contracted, as shown in FIGURE
4A", the floating seal should be near its lowermost normal
position due to the lubricant flowing into the space 283 at
its top side. If the damper is fully extended, as shown in
Figure 4A', the floatiny seal should be near its uppermost
normal position, due to the lubricant flowing away from
space 283 at its top side. If the damper is in its neutral
position, the seal should be in a normal median position, as
shown schematically in Figure 1.
'
If the seals leak so that there is a loss of
lubricant from the chamber, the floating seal will be near
its uppermost position, or at least above its normal posi-
~57~ 9
tion. If the seals leak in a ~anner that intrusion of wellbore ~luid increases the fluid in the charnber, the floa-ting
seal will be at its lowermost position or at least below its
normal position.
O course in both conditions, insuficient and
excess li~lid in the chamber, the problem may not be with
the seals but rather be one of initial insuficient or
excessive filling o the chamber with lubricant. Whatever
the problem is, it can be corrected.
- Springs -
Referring again to FIGURE 1, and more particularly
to FIGURE 3, each of spring stacks 167, 169 comprises a
plurality of Belle~ille spring washers 321 positioned with
their cones pointing up, interleaved with a plurality of
Belleville spring washers 323 disposed with their cones
pointing down. This mode of stacking places the spring
washers in series. The more spring washers in series, the
greater the damper deflection for any given axial load. The
use of two spring stacks in parallel reduces the stress in
each Belleville spring washers for any given deflection of
the damper. If reguired, additional stacks of spring washers
beyond two stacks e.g. three, four or more stacks may be
parallel to keep the stress in each Belleville spring washer
below the elastic limit or below the endurance limit or any
other desired limit.
; It may be noted that merely making the spring
washers thicker or stacking some of the Belleville spring
washers in each unit in parallel, that is with their cones
_ ~3 ~
1157~ 9
pointing in the same direction, though reducing some of the
stresses in the spring washers, will not change the local-
ized stresses over the areas o contact between adjacent
oppositely pointed spring washers, which stress may b~ very
high due to the nearly line contact between~such spring
washers and will not change the localiæed tensile stresses
on the faces of the spring washers opposite to their areas
of contact, and furthermore, in the case of springs paral-
leled within the stack, will cause sliding friction between
the springs thus paralleled. Therefore paralleling several
stacks may be necessary.
.Due to manufact~ring tolerances, the length of a
spring stack of a preselected number of spring washers may
not exactly fit the cavity between mandrel and barrel. In
such case, as shown only in Figure 1, one or more flat
washers or shims 280, 282, 284 may be employed to achieve
the~ desired.fit. Alternatively, pressure rings 164, 165,
174, 183 (Figure 3) of varying thickness may be employed.
,
- Variable Moment Belleville Springs -
3ue to the small seal, Figure 1 shows spring
stac~s 167, 169 .to be composed of ordinary Belleville spring
washers, i.e. with cross sections having straight sides.
Although such springs may be employed while obtaining some
of the advantages of the invention, and may even be con-
structed to provide a variable modulus as set forth, e.g. at
pp. 155-157 of Mechanical Springs, by A. M. Wahl, Second
Edition, published by McGraw Hill Book Company, 1963, it is
preferred to use variable moment arm Belleville springs of
~LS71!3 ~9
the general type d:isclosed ln the aforernentioned Migny
patent. Furthermore it ls pxeferred to use a particular
form of vaxiable moment arm Belleville spring, herein termed
a roller Belleville spring, in which there is pure rolling
between adjacent washe~s as they are loaded and unloaded, as
next described.
- Roller Belleville Springs -
The Belleville spring washers 321, 323 are identi-
cal, merely being positioned oppositely during assembly.
Such a spring washers is shown to a larger and more precise
scale in FIGURE 7. As there shown, the radius R' - R" of
the outwardly and inwardly tilted faces of the spring washer
is 10-11/64 lnches. The outer diameter of the spring washer
is typically ~.827 inches. The spring washer thickness is
0.250 inch and the cone angle is 3.75 degrees measured at
the part of the spring washer midway between its inner and
outer peripheries. It will be noted that the center 0" of
the radius R" for the outwardly facing face (the left hand
face in Figure 7) of the washer lies substantially on a line
through the inner peripheral edge of such face parallel to
the cone axis C of the washer. When the washer is stacked
with others and assembled in the damper, the slight preload
in the neutral position will cause the center of such radius
R' to be exactly on said line.
Similarly it will be noted that the center 0' of
radius R' for ~he inwardly facing face (the right hand face
of Figure 7) of the washer lies substantially on a line
through the outer peripheral edge of such face parallel to
the cone axis of the washer. When the washer is stacked
with others
~ 57 ~9
and assembled in the darnper, the slight preload in the
neutral position will cause the center of such radius R" to
be exactly on said line.
With O' and o" so located for the neutral position
of the d~mper, and R 'being e~lal to R~", the contacting
areas of the washers of the inner and out~r peripheries of
the stack will, ~e tangent; when the damper is deflected the
contacting areas of the washer move closer together and
remain tangent, the washers rolling upon each other without
sliding. It will be noted -that when the contacting areas of
the washer move toyether so as to be in vertical alignment,
that is, so that they are equidistant from the axis of the
washers, the moment arm becomes zero and the spring has gone
solid. During the motion ~rom unloaded condition to the
solid condition, only the inner portions of the washer faces
that are in contact nearest the washer axis are engaged and
only the outer portions of the washer faces that are in
contact farthest from the washer axis are engaged, such
portions being the functional portions of the surfaces. The
non-functioning portions of the washer surfaces never engage
any other surface.
With the foregoing background, the conditions for
pure rolling of the washer may be summarized as follows:
~1) The contacting areas of the faces of the washers
are tangent, to avoid pivoting.
(2) The washers are identical, i.e. of like size,
thickness, dish (cone angle), and facial curva-
_ ~6 -
~157B `~9
ture, so tllat they will havc equal angles of
cleflection ~m-l the same position oE thelr neutral
axes, as requirecl to avoid slippage.
(3) To avoid slicling, the curves of tlle functional
portions of the two opposite faces of each washer
should be thc same, i.e. one curve would be the
double reflection of the other about a medial cone
and a cone perpendicular thereto (cor~lementary
therewith).
(4) The curves of the functional portions of the cross
sections of the surfaces of the washers must be
continuously convex, to avoid bridging.
- Spring Clearance and Guidance -
Referring once more to Figure 7 the mlnimum diame-
ter of the inner periphery of the washer is 03.543 ~ 0.005
inches~ which is enough larger than the diameter d of the
outer periphery of mandrel 33 (Figure 5) that the spring
washers can freely move up and down axially relative to the
mandrel even when the damper is extended or contrac~ed
;~ 20 ~ causing the washers to be compressed ~flattened). If any
part of the spring stack moves laterally, the mandrel will
limit the movement, one or re of the spring washers making
line contact with the side o~ the mandrel. The inner corners
~ of the cross-section of the spring washer are more fully
; rounded to avoid cocking on the mandrel during assembly and
to reduce stress concentration.
- 47 -
:L~S78~9
The outer periphery of the spring washers is
considerably smaller than the ilmer diameter D (FIG. 5) of
the barrel, so that the washers can expand circumferen-tially
when they are compressed (flattened~, as the damper extends
or contracts, and still remain out of contact with the
barrel. Also, there is left plenty of room for lu~ricating
oil 91 to move past the washers.
If desired, the washers could more nearly make a
close fit with the barrel and, at the same time, if so
desired, less nearly make a close fit with the mandrel,
relying on the barrel to limit lateral travel of the springs.
. The minimum radial clearance betweenthe inner and
outer peripheries of the washers, and the adjacent outer
periphery of the mandrel and inner periphery of ~he barrel
re~uired to accommodate for change in the spring washer
diameters when flattened is only a few thousandths of an
inch, which is less than that required by manufacturing
tolerances.
- Variable Spring Modulus -
The spring stacks, when assembled in their respec-
tive pockets, are slightly compressed even when the damper
is neither contracted nor extended, but only just enough to
keep them from being loose in the~ pockets. In this condi-
tion, as shown in.~ S~ L 7A/ the spring washers are
in contact over circular areas near their inner and outer
peripheries, each spring washer contacting either the inner
or outer periphery of the spring washer above and the oppo-
~L~S7~
site (outex or inner) periphery of the spriny ~1asher below.This is the unloaded condition of -the damper.
When the spring stacks are compressed, upon either
contraction or extension of the damper, the circular areas
of contact between the spring washers move away from the
peripheries toward each other, i.e. towards the mid widths
of the washers, as shown in Figure 3. This results in a
reduction of the leverage of the axial forces on the spring
in their action to flatten the washers, causing the spring
rate (force/deflection ratio) to increase. Roller Belleville
springs thus have a pronounced variable spring modulus.
Referring now to FIGURE 8, there aré shown a curve
A plotting spring force versus deflection, and also a curve
X plotting spring rate versus deflection. Curve X shows a
very low spring rate of the order of 4,000 to 10,000 lb./in.
at moderate values of spring deflection, reflecting the low
slope of the nearly straight line portion of the spring
force deflection curve A below 10,000 lb. At a deflection
of l.O inch, corresponding to a spring force of 12,500 lb.,
the spring rate is stilI less than 30,000 lb./in. Even at a
deflection of 1.3 inches, corresponding to a spring force of
30,000 lb., the spxing rate is but 85,000 lb./in. Only as
the deflection approaches the travel limit of 1.5 inches
does the spring rate exceed 100,000 lb./in. to bring the
damper travel to a cushioned stop.
~1 57 8
- Double Action -
FIGURE 8, curve A, also shows that for negative
loads, i.e. loads tending to extend the damper, the load
deflection curve is reflected about the ordinate. In other
words, the same load deflection curve, except with negative
deflections, applies. That is because the resilient means
is being stressed regardless of whether the damper is being
contracted or extended. In fac~, in the preferred embodi-
ment the same spring means is strained in the same way
(~lattened) regardless of whether the damper is contracted
or extended.
Balanced Load Drilling
Referring now to Figure 9, there is shown curve A,
which is the same as curve A of Figure 8, plotting spring
force as a function of spring deflection. Figure 9 also
shows a curve B plotting spring force as a function of
decreasing spring flexibility, that is, the abscissae scale
of curve B has its origin at a flexibility of 25 inches per
100,000 pounds, with decreasing flexibility proceeding away
from the origin. Also, flexibility is plotted as positive
both to the left and to the right of the origin; this ac-
counts for the fact that the flexibility curve incllldes a
portion in the lower left hand ~uadrant rather than in the
upper left hand quadrant. Flexibility is the reciprocal of
spring rate; therefore curve B is closely related to curve X
of Figure ~. However spring rate curve X is plotted against
deflection whereas curve B plots sprlng force against flexi-
bility. Note further that in curve X, spring rate is plotted
_ So _
~ 57 ~9
as ordinates, whereas in curve B flexibility is plotted as
abscissae. In fact in Figure 8, one should consider the
abscissae, the deflection, as the independent variable,
reflecting the ~act that the bit moves up and down as the
contour of the bottom changes, to some extent regardless of
what force is imposed on the bit, whereas Figure 9 i9 best
appreciated viewing the ordinates, the spring force, as the
independent variable, reflecting the information known to
the driller, i.e. the static load on the damper.
It may be noted here that although the static load
on the damper is -the dlfference between the drilling weight
and the pump apart force, the load on the bit is the drill~
ing weight, unaffected by the pump apart force. Although
the pump apart force acts down on the bit, it also acts
upwardly on the swivel to relieve the drawworks of some of
the drill string weight. Since drilling weight is calcu-
lated on the basis of weight of drill string less line
tension, the upward pump apart force, which actually further
reduces the unsuspended weight of the drill string, equals
the downward pump apart force acting to increase the force
on the ~it over that which is due to the unsuspended weight
of the drill string. In other words, the pump apart force
is neglected twice with opposite effect.
Figure 9, in the upper left hand quadrant, includes
a table showing typical drilling conditions. The items in
the table, and elsewhere in Figure 9, that are marked with a
check mark, correspond to a near balanced load condition,
CONDI~ION I, including an average drilling weight of 45,000
pounds and an average pump pressure of 1,500 psi correspond-
_S i--
1~S7~3 ~9
ing to a pump apart force of 44,000 pounds. The itemsmarked with an asterisk correspond to an extreme condition,
CONDITION II, wherein the pump apart force dominates, the
pump pressure of 2,000 psi producing a pump apart force o~
59,O00 pounds compared to a dri:lling weight of only 30,000
pounds, a difference of 29,000 pounds acting to expand or
extend the damper. The items marked with a double dagger
correspond to an extreme condit:ion, CONDITION III, wherein
the drilling weight dominates, the pump pressure of only
1,000 psi producing a pump apart force of only 29,000 pounds
compared to a drilling weight of 60,000 pounds, a difference
of 31,000 pounds acting to compress or contract the damper.
Having noted the parameters in the upper left hand
quadrant of Figure 9, one may refer to the scale of pump
pressures at the lower left hand side of Figure 9. There,
selecting a pump pressure of 1,500 psi, marked with a check
mark, and following the heavy line in the direction of the
arrows, one finds that this pressure corresponds to 44,000
pounds of pump apart force, a negative force, i.e. one
expanding the damper, to which one adds a posi-tive force o~
45,000 pounds of drilling wei~ht to provide a net force of
1, oao pounds contracting the damper, at which load the
flexibility of the damper is a maximum~ i.e. 25 inches per
100,000 pounds. This is CONDITION I.
Further exploring Figure 9, one may start at the
lower left at a pump pressure of 2,000 pounds marked with an
asterisk, and following the dotted line one first notes that
this pump pressure produces a pump apart force of minus
59,000 pounds. To this is then added a drilling weight of
7~349
30,000 pounds leaving a net negati~e force of 29,000 pounds
expanding the damper, at which point the ~lexihility of the
damper is 0.85 inch per 100,000 pounds. ~his is CONDI-
TION I I .
Condition III may also be traced on Figure 9,starting at the lower left with a pump pressure o 1,000
psi, marked with double dagger. Followi.ng the broken line
one finds that 1,000 psi pressure corresponds to a pump
apart force of minus 29,000 pounds. Adding a drilling
weight of 60,000 pounds creates a net contractive force on
the damper of 31,000 pounds, which corresponds to a flexi-
bility of 0.80 inch per 100,000 pounds, substantially the
same as for CONDITION II.
Assuming a case wherein the pump apart effect
balances the unsuspended weight of the drill string ~drill-
ing weight), identified as CONDITION I in Figure 9, there is
no static load on the damper spring and the damper will
operate about the zero load, zero deflection point, the
origin of the FIGURE 9 load-deflection curve, which is the
neutral position as previously defined. The damper will
then be very soft and very little motion will be transferred
to the drill string. This is in contrast with single acting
variable modulus dampers in which under balanced load condi-
tion alternate half cycles of vibxation would cause engage-
ment of the travel stops, thereby losing the benefit of the
low spring rate on alternate haIf cycles of damper vibration.
.
Consider next the case of drilling with unbalanced
load. As the unbalance increases, the flexibility at the
_ S~--
static deflection point decreases~ However, at a load
unbalance of plus or minus 10,000 lb., the flexibility is
still over 4 inches per 100,000 lb., which is very high, and
the deflection is about 0.95 in. which leaves 0.55 inch of
tra~rel to the nearest travel stop.
It may be no-ted here that the drop in flexibility
from 25 inches per 100,000 pounds under balanced load condi-
tions to 4 in./100,000 lb. at 10,000 pounds unbalance appears
to be major. However the amplitude of transmitted vibration
to impressed vibration is actually more nearly a function of
the reciprocal of flexibility, i.e. of spring rate, and as
appears from Figure 8, at a deflection of 0.95 in. the
spring rate is only about 23,000 pound/in., which is still
quite low.
~ If the drilling weight exceeds the pump apart
;~ effect, or vice versa, by thirty thousand pounds, the condi-
tions are those identlfied on Figure 9 as CONDITION II and
CONDITION III. Conditions II and III represent extreme
conditions of load unbalance which are met in practice per-
haps only about 5 per cent of the time. However even under
these conditions although the action of the damper is not as
good as under balanced load, the damper, having a 1exibil-
ity of 0.8 or more and a travel of 0.2 in to the nearest
stop, is still soft enough to dampen substantially transmis-
sion of vibration to the drill string. The damper therefore
may be used with varying degrees of effectiveness over a
typical range of drilling weights in the range from 30,000
pounds to 60,000 pounds and pump apart forces in the range
from 30,000 pounds to 60,000 pounds corresponding to a 6-1/8
l~LS~ 9
inch diameter seal area (e.g. 8 inch diameter damper) with
pump pressure in the range of 1,000 psi to 2,000 psi.
-Stroke-
Assuming the drill string above the damper to be
at rest vertically, the damper needs to have sufficient
contractive and extensive stroke to allow for the maximum
anticipated rise and fall of the drill bit without having
the damper become inoperative by the travel limit stops
becoming engaged or the springs reaching their limit of
deformation (going solid i.e. flat in the case of Belleville
springs). It will be seen from FIGURE 8 that the damper has
a working range of plus or minus one inch deflection with a
very low spring rate when the pump apart effect balances the
drilling weight. Even at the extreme condition of a thirty
thousand pound difference (plus or minus) between drilling
weight and pump apart effect, there is still available a
travel of about .2 inch in the direction toward the nearest
travel limit stop before the spring rate of the spring
stacks becomes exceedingly high.
~; If one is willing to accept a higher spring rate
at balanced load, the available travel to the nearest stop
for any given unbalanced load and the spring rate at that
load can both~be enhanced by employing additional stacks of
springs in parallel. Referring to Figure 9A, curve A shows
the load deflection curve for the previously described
apparatus. Curve Al shows the result of employing four
stacks of springs in parallel (twice as many as for Curve A).
It is seen that although the spring rate for balanced load
SS ~
49
is doubled (twice -the slope), it is still very low, and at
30,000 lb. static load the deflection is only 1.1 inch,
].eaving 0.4 inch travel to the nearest stop (compared with
0.2 in. for curve A) and the spring rate is lower (lower
slope) than fo~ curve A.
Curve Al also shows that at 60,000 lb. static load
(twice the limit of the working range for the da~per of
Curve A) the deflection is only 1.3 inches, the same as for
a 30,000 lb. load in the case of the curve A damper (although
the spring rate is greater for curve Al at 60,000 lb. than
for curve A at 30,0Q0 lb.
Summarizins, by puttlng more variable modulus
springs in parallel, one can not only reduce the deflection,
but also reduce the spring rate at the same 30,000 lb.
static load, or increase the static load for the same 1.3
inch deflection.
I~ an overall longer stro~e is required, the
lengths of the spring stacks can be increased. For example,
referring now to curve A~ of Figure 9, by doubling the
length of the spring stacks, the low spring rate working
range would become plus or minus two inches, and there would
be a travel of about .4 inch to the nearest travel stop when
static load is unbalanced by 30,000 pounds.
Lengthening the spring will also lower the spring
rate in the middle of the range, e.g. at the neutral posi-
tion. Therefoxe if it is desired to increase the working
range of unbalanced load-as well as increase the stroke,
5~ ~
~ ~. 57 'B L~9
without sacrificing flexibility at any load, additional
stacks of springs in parallel may be employed. For example,
by both doubling t~e lengths of the springs and employing
twice as many in parallel as shown in curve A3, the same low
spring rate at midrange as ~or curve A w.ill be achieved, and
the unbalanced load working range for the same maximum
spring rate within the range will be increased to plus or
minus 60,000 lb. with a travel limit o 0.4 inches to the
nearest stop.
It is to be particularly noted from curve A2 that
by doubling the spring length and the number of stacks in
parallel, the travel to the nearest stop at plus or minus
30,000 pounds would be increased to 1.08 inches and the
spring rate would be only about 25,000 lb/in. This improved
effect achieved with the seriating and paralleling of vari-
able modulus springs is in contrast with that achieved with
constant modulus springs where only the travel to nearest
stop would be increased without any change in spring modu-
lus.
.
-COMPARISON-
The subject construction with a stroke of plus or
minus 1.5 inches and a spring rate of 30,000 lb./in. or less
over a range o~ plus or minus one inch when operating under
balanced load conditions, is to be compared with the result
that would be obtained with a constant spring rate and with
single acting spring means.
S 7--
1~L57f~ 9
First of al1 consider the situ~tion with a con-
stant spring rate sinyle acting ~pring means having a spring
rate of 4,000 lb. per inch, corresponding to the spring rate
of the presen~ construction at zero deflectlon. Upon imposi-
tion of a static load of ~hirty thousand pounds, the deflec-
tion would be 7.5 inches, which is way beyond the range of
travel of the ass~ed situation. To accomodate such a
stroke, the seals, spline, and c~ide bearings would all have
to be lengthened, as well as providing a spring of such a
stroke. The question arises lf this p~oble~ could he over-
come merely by changing the position of the travel limit
stops. The answer is simply, no. If the stops were set to
limit deflection to plus or minus 1.5 inch, upon imposition
o~ an unbalanced static load of only 6,000 pounds the damper
would be in engagement with one stop and therefore inopera-
tive on alternate half cycles.
At this point one may introduce the concept of the
load carrying capability of a spring. This is the maximum
load which a spring can carry without going solid, or other-
wise expressed, the maximum load which a spring can carry
and still function as a spring, i.e. as a device in which
the ratio of load to deflection per unit length of spring is
less than the modulus of elasticity of the material of which
the spring is made. In the instant case, unless the spring
of the spring means has a load carrying capacity of thirty
thousand pounds, it will not be effective as a spring upon
imposition of a 30~000 pound load, no matter where the
travel limit stops are placed. If a spring is soft, it
likely will have a low load carrying capacity.
_ 5~
~S~8'~g
Consider next a damper with a constan-t spring rate
of 30,000 lb./in., near the maximum rate for the damper of
the present invention, when operating under balanced load
conditions. Under a static load of 30,000 lb., the spring
deflection would be one inch, leaving only a half inch of
travel to reach the nearest travel limit stop. Since a one
inch deflection is to be expected according to the original
assumptions, such a damper would strike its travel limit
stop once during each cycle of vibration, there being no
increasing modulus as the stop is approached to cushion the
end of the travel.
To provide a one inch travel to the nearest stop
when operating with a 30,000 lb. static load, the spring
rate would have to be such that the static deflection would
be only one-half inch (1.5-1.0 = 0.5). In other words, a
constant spring rate of 30,000jO.5 = 60,000 lb./in. would be
required. This is fifteen times as stiff as the zero deflec~
tion spring rate of the previously described construction
embodying the lnvention. In addition, should there be any
abnormal deflection of a damper with the assumed 60,000
lb./in. constant spring rate, the damper would strike its
travel limit stop. In contrast, the spring means of the
present construction, being of increasing spring rate as the~
deflection approaches the limits, will still effect a cush-
ioned stop upon abnormal deflection, even though the cushion-
ing will be less than under normal conditions.
Consider next the case of a known damper employing
a variable modulus spring but with single action. Such a
damper operating under balanced load would strike its travel
limit stop once during each vibration.
5C/ _
~571~349
CONST~NT BIT LO~D DRILLING
The type of drilling contemplated by the present
invention may be further understood by referring once more
to FIGURE 8. It will be seen f:rom curve X that for deflec~
tions between plus and minus 0.5 inches the spring rate is
nearly constant and guite low. Therefore, if drilling is
conducted in such a manner that the pump orce balances the
drilling weight so thak the static deflection of the damper
is zero, the damper will allow the drill bit to move up and
down following the contour of the bottom of the hole under
the constant downward ~orce of the pump apart force with
very little force variation transmitted to the drill string
through the damper. Since the bit will remain in contact
with the bottom of the hole with the desired~force on
bottom, drilling should proceed in a most efficient manner
and at the same time there~will be:minimum wear and tear on
the drill string.
~: : MO~IFICATION
Although according to the preferred embodiment of
the invention roller Belleville springs are employed, vari- :
able modulus spring means other than roller Belleville
:: springs may also be used.
:~ As an example of the latter construction, FIGURE
illustrates a form of spring stack to be disposed in a
damper pocket, similar to the pockets 162, 163 in the pre-
viously described embodiment. The stack comprises ovoid
_ (~0_
., '
~578~
cross-section elastomer quoi-ts 301 each sandwiched between
pairs of L-shaped cross-section steel washers 303, 305 which
slide Ereely in the pockets between mandrel and barrel. The
washers are in peripheral engagement, placlng the quoits in
parallel. Groups of paralleled quoits, e.y. group of three
quoits, are separa-ted by s-teel spacers 307 placing the
groups in series. With this arrangement, by selecting the
number of quoits in a group, the total number of groups, and
~he cross-sectional shape of the quoits, most any stress
limit on the quoits and travel range for the damper can be
obtained. As compared with roller Belleville springs, a
softer, no load spring rate will be typical, with a sharply
increasing rate as the elastomer quoits deform and the metal
washers come closer together and the elastomer fills the
voids 309, 311 present at the inner and outer peripheries of
the quoits when unloaded.
Though the rubber quoit sandwich construction has
the advantages of greater softness, a disadvantage is the
fact that rubber cannot be used in high temperature wells.
Also, rubber has considerable hysteresis or internal fri~-
tion which may reduce the life of the spring means.
While a preferred embodiment of the invention and
a modification thereof have been shown and described, many
further modifications can be made by one skilled in the art
without departing from the spirit of the invention.
.
