CA1157849A - Drill string splined resilient tubular telescopic joint for balanced load drilling of deep holes - Google Patents

Abstract

DRILL STRING SPLINED RESILIENT TUBULAR TELESCOPIC
JOINT FOR BALANCED LOAD DRILLING OF DEEP HOLES
Abstract of the Disclosure A drill string splined resilient tubular tele-scopic joint for balanced load deep well drilling comprises a double acting damper having a very low spring rate upon both extension and contraction from the zero deflection condition. Preferably the spring means itself is a double acting compression spring means wherein the same spring means is compressed whether the joint is extended or con-tracted. The damper has a like low spring rate over a considerable range of deflection, both upon extension and contraction of the joint, but a gradually then rapidly increased spring rate upon approaching the travel limits in each direction. Stacks of spring rings are employed for the spring means, the rings being either shaped elastomer-metal sandwiches or, preferably, roller Belleville springs. The spline and spring means are disposed in an annular chamber formed by mandrel and barrel members constituting the tele-scopic joint. The spring rings make only such line contact with one of the telescoping member as is required for guid-ance therefrom, and no contact with the other member. The chamber containing the spring means, and also containing the spline means, is filled with lubricant, the chamber being sealed with a pressure seal at its lower end and an inverted floating seal at its upper end. Magnetic and electrical means are provided to check for the presence and condition of the lubricant. To increase load capacity the spring means is made of a number of components acting in parallel.

Description

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.
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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).

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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."

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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."

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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.

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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.

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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.
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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~

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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.

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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;
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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;
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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;

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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.

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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.
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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 .

~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 ~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.

: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 ~

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.

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.

.

Claims (20)

THE EMBODIMENTS OF THE INVENTION IN WHICH AN EXCLUSIVE PROPERTY OR
PRIVILEGE IS CLAIMED ARE DEFINED AS FOLLOWS:

1. A drill string vibration damper comprising a splined tubular telescopic joint and resilient means urging the joint to a neutral posi-tion 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 increasing spring modulus upon departures from said neutral position in both directions.

2. Damper according to claim 1, said resilient means being strained upon both extension and contraction of the joint from a neutral position and having a like variable spring modulus upon both extension and contraction with a lowest modulus at the neutral position and nearly as low like moduluses at equal departures from neutral position to both sides of the neutral position over a range of such departures and like gradually and then sharply higher moduluses upon increasing equal departures beyond said range upon further extension and contraction of the damper.

3. Damper according to claim 1, said joint comprising telescopically disposed mandrel and barrel members, connection means at one end of each member for making connection with a rotary drill string component, means for transmitting torque between said elements while allowing rela-tive axial motion therebetween, said resilient means being separate from the last mentioned means resil-iently transmitting axial forces between said elements upon relative axial motion of said members m both directions from said certain position, said mandrel and barrel members being tubular providing a flow passage therethrough, sliding pressure seal means between said members sealing said flow pass-age against fluid flow from said passage outwardly through said tool between said elements and vice versa, floating seal means between said members forming with said pressure seal means an annular chamber between said members, and lubricant in said chamber, said resilient means being disposed in said chamber, said resilient means being at all times, as said members move axially relative to each other, radially spaced from one of said members and freely movable axially relative to the other said members.

4. Damper according to claim 1, said joint comprising an inner tube adapted at one end for connection to a drill string component, an outer tube adapted at one end for connection to a drill string component, the other end of the inner tube being telescopically disposed within the other end of the outer tube, axially spaced means sealing between said inner and outer tubes defining a volume between said tubes, which volume contains lubricant, and disposed in said volume torque transmission means to transmit torque between said tubes and said resilient means which serves to transmit axial force between said tubes, said resilient means including in parallel a plurality of groups of series stacked spring rings, said resilient means being strained upon both separation and approach of said one end of said tubes from said certain position and having a variable spring modulus increasing upon wide departure of said tubes from neutral position.

5. Damper according to claim 1, said resilient means comprising series stacked spring rings.

6. Damper according to claim 1, said resilient means being disposed in a bath of lubricant, and an electric conductor extending through said joint from the exterior thereof into contact with said lubricant to facilitate measurement of the electric resistivity of the lubricant.

7. Damper according to claim 1, said telescopic joint including a mandrel and a barrel and axially spaced seal means sealing between the man-drel and barrel defining volume between said mandrel and barrel adapted to receive lubricating fluid, said resilient means being disposed in said volume, one of said seal means being sliding pressure seal means and the other of said seal means being floating seal means, and magnet means carried by said floating seal means to facilitate determination of the position of the floating seal means.

8. Damper according to claim 7, said barrel including a wash pipe extending inside the end of said mandrel, said floating seal means com-prising annular means around the outside of the wash pipe and inside the mandrel, said magnet means comprising a permanent magnet carried by said annular means.

9. Damper according to claim 1, said telescopic joint including a lower mandrel member and an upper barrel member therearound and axially spaced annular seal means sealing between the mandrel and barrel defining a volume between said mandrel and barrel adapted to receive fluid lubricant, said resilient means being disposed in said volume, said seal means having lower and upper faces disposed with their upper faces to be adjacent lubricant when said volume is filled with lubricant.

10. Damper according to claim 1, said resilient means including a plurality of paralleled groups of series stacked spring rings.

11. Damper according to claim 1, said resilient means including a plurality of series groups of parallel stacked spring rings.

12. Damper according to claim 1, said resilient means including a plur-ality of parallel groups of series subgroups of parallel stacked spring rings.

13. Damper according to claim 1, said resilient means comprising a plura-lity of elastomer metal sandwiches, each sandwich including upper and lower 64 .

metal washers with an ovoid cross-section elastomer ring therebetween, each upper washer having a depending flange from one of its inner and outer peripheries and said lower washer having an upstanding flange at its opposite periphery, said flanges overlapping said washers, and spacer washers disposed between groups of said sandwiches, the inner peripheral flanges engaging each other and the outer peripheral flanges engaging each other within each group of said washers, said sandwiches in each group not being in parallel with respect to transmission of force, said spacer washers having inner and outer peripheries clear of said flanges as said damper extends and contracts so that said groups are in series with respect to transmission of force.

14. Damper according to claim 2, said joint comprising axially movable members with axially spaced seal means between said members defining a lubricant chamber housing said resilient means, and an electrical con-ductor extending from the exterior of said joint to said chamber and electrically insulated from said joint to facilitate measurement of the electric resistivity of the lubricant.

15. Damper according to claim 2, said joint comprising axially movable members with axially spaced seal means between said members defining a lubricant chamber housing said resilient means, both of said seal means having upper surfaces positioned to contact lubricant in said chamber and lower surfaces positioned to contact drilling fluid.

16. Damper according to claim 2, said joint comprising axially movable members with axially spaced seal means between said members defining a lubricant chamber housing said resilient means, the outer one of said members having a wash pipe extending within the other of said members, one of said seal means being a floating seal disposed between said wash pipe and said other member, said floating seal means being provided with a permanent magnet to indicate its position through said wash pipe, said wash pipe being paramagnetic.

17. Damper according to claim 2, said joint comprising axially movable members with axially spaced seal means between said members defining a lubri-cated chamber housing said resilient means, said resilient means comprising a plurality of elastomer metal sandwiches, each sandwich including upper and lower metal washers with an ovoid cross-section elastomer ring there-between, each upper washer having a depending flange from one of its inner and outer peripheries and said lower washer having an upstanding flange at its opposite periphery, said flanges overlapping said washers and spacer washers disposed between groups of said sandwiches, the inner peripheral flanges engaging each other and the outer peripheral flanges engaging each other within each group of said washers so that said sandwiches in each group are in parallel with respect to the transmission of force, said spacer washers having inner and outer peripheries clear of the said flanges as said damper extends and contracts so that said groups are in series with respect to the transmission of force.

18. Damper according to claim 1, said resilient means comprising a first variable Hooke's modulus spring means and a second variable Hooke's modulus spring means, said variable Hooke's modulus spring means being separate and in parallel.

19. Damper according to claim 18, each said variable modulus spring means comprising a stack of seriate ring elements, the ring elements of the stack of the first variable modulus spring means being out of contact with the ring elements of the stack of the second variable modulus spring means.

20. Damper according to claim 19, each of said stacks comprising a plurality of resilient plastic material members each sandwiched between metal members.