Note: Descriptions are shown in the official language in which they were submitted.
~257~3~65
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The present invention relates to improvements in the
sealing arrangements ~or a bearing, or a combined bearing and
gear-reduction assembly, of a down-hole driven drill apparatus.
~ACKGROUND OF THE INVENTION
Down-hole driven drill assemblies include a drill string
that is located within a well bore. A bit is located at the end
of the string. Positioned near the bottom of the drill string,
and located within the stationary housing of the drill string, is
a down-hole motor. In any drill assembly of the type used, for
example, in drilling for hydrocarbons or for geother~al sources,
the drilling string is hollow. Drilling fluid or mud is pumped
down the hollow interior of the string, outwardly across the face
of the bit, and up the annular space between the exterior of the
drilling string and the interior of the well bore. A down-hole
motor is powered by means of the drilling mud or fluid which, in
moving through the motor, rotates or drives an inner vertical
shaft or mandrel within the drill string which, in turn, drives
the drill bit that is located at the end of the drill string and
is secured to the mandrel. The mandrel is rotatably associated
with a bearing assembly that is provided between the mandrel and
the stationary outer casing of the drill string. The bearing
assembly transmits radial and axial forces between the stationary
casing and the driven mandrel.
~ own-hole motors are generally one of two types, namely,
either a positive displacement motor or a turbine motor.
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Positive Displacement Down-Hole Motors
In a positive displacement motor, whether 1-2 lobe or
multi-lobe, a given volume of fluid pumped through the motor will
produce a certain number of revolutions of the rotor. This is not
affected by mud weight, torque loading (within the motors oper-
ating range), bottom hole pressure or pressure drop across the
bit. This type of motor in the past was used mainly for direc-
tional drilling but is now also used frequently for straight hole
drilling.
A complete motor assembly consists of 4 main components,
namely, starting from the top of the motor: a bypass valve; the
motor section (rotor & stator); universal joints and housing; and
the bearing assembly.
~pass Valve
The bypass valve located at the top of the tool, allows
drilling fluid to bypass the motor and fill the pipe when tripping
in the hole and allows the pipe to drain when tripping out of the
hole. This valve automatically closes as soon as the pump is
started. Without this valve the motor will run in reverse when
tripping in the hole, and wet trip out of the hole.
Motor Section (1 to 2 Lobe)
_
The motor section consists of 2 parts, a rubber moulded
stator with a spiraled, oblony interior cross sectional profile,
and a steel rotor with a spiraled round cross sectional profile.
Rotation is caused when fluid is forced under pressure into the
cavities which are formed between the rotor and stator. Because
of the rubber in the stator it has a temperature limit of 300~F.
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The use of oil ~ased drilling fluids will cause premature failure
of the rubber in the stator. ~ormal life of a stator is between
300 and 600 hours before relining.
The rotor is normally chrome plated and will require
replating in 500-800 hours. If used in salt based drilling
fluids, rechroming will be required much sooner. Most of these
motors are either 3 or 4 stage. Each stage consists of 1 spiral
in the stator and 2 on the rotor. ~ach stage requires a pressure
drop of approximately 12~ PSI for maximum power. The rota~ional
speed of the motors can be altered during manufacturing within a
practical limit by increasing the lead of the spiral to lower the
speed and decreasing the lead to increase the rotational speed.
Torque will change very little with a change in lead as long as
overall length of motor is constant.
Drive Shaft & Housing
The rotor rotates in the stator in a eccentric motion,
so a drive shaft with a universal joint is required to convert
this to concentric rotation. This is a problem area with all
positive displacement motors, and a major cause of failure.
Universal joints presently being used are grease packed, with a
rubber hose clamped at each end covering them. I-f the hose punc-
tures or leaks in any way, the universal joint fills with drilling
mud, causing extreme wear and eventual failure. Oil based drill-
ing muds wil] cause the hose to deteriorate. The drive shaft is
housed in a drive shaft housing, which will quite often have the
threaded connection at one end of the housing cut at an angle to
the axes of the housing when used for directional drilling.
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Bearing Assembly
Most motors presently on the market have mud lubricated
bearing assemblies. A flow restrictor allows about 10-15% of the
mud to pass through the bearing assembly. Both ball thrust
bearings and angular contact bearings are used for axial loads.
Friction bearings are used for radial loads and these are usually
rubber lined marine or hard metal sleeves and bushings.
Positive Displacement Multi-Lobe Motors
To decrease rotational speed and increase torque, multi-
lobe motors have recently been developed. In this design therotor has one less lobe ~han the stator. This geometric configur-
ation acts as a gear reduction giving a lower speed and higher
torque. Multi-lobe motors range from 2-3 lobe, to as high as
10-11 lobe. Multi-lobe motors are not as efficient as 1-2 lobe
motors and the efficiency decreases with an increase in the number
of lobes. Increasing the number of lobes increases torque, while
decreasing rotational speed. The increase in torque is not in-
versely proportional to the decrease in rotational speed (for a
given pressure drop and flow rate), causing a decrease in power.
The operational life of multi-lobe motors is normally less than
that of 1-2 lobe motors.
Turbine Down-Hole Motors
Turbines are normally used for straight hole drilling,
and are almost exclusively used with drag bits of either diamond
or polycrystalline diamond coated carbide bits. When turbines are
used with tri-cone bits, the life expectancy of the bits is very
short because of their high rotational speed. Turbines require
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~higher flow rates and higher pressure drops across the rotors and
stators than positive displacement motors. The efficiency of a
turbine is approximately 50%, while that of a conventional
1-2 lobe positive displacement motor is approximately 80%.
Because of the high volumes and pressures required by the tur-
bines, they are restricted to drilling rigs with capable pumps.
Their length causes some limitations in directional drilling.
Turbines have the advantage over positive displacement
motors in that they can withstand higher bottom ho]e temperatures
and are not adversely affected by oil based drilling fluids.
The rotational speed of a turbine motor is directly
proportional to the flow rate of the fluid pumped through the
turbine and inversely proportional to the torque load applied.
Although flow rate through the turbine is usually known, rota-
tional speed is not because accurate torque reading at the drill
bit is difficult to obtain. The torque output of a turbine is
maximum at stall, and zero torque at maximum or runaway speed.
Maximum power is generated at a speed approximately half way
between stall and runaway. The rotary speed of turbines, within
limits, can be determined by blade design. However, turbines are
not practical below about 400 r.p.m.
The pressure drop across the rotors and stators and
across the bit create a force extending the output shaft from the
turbine housing. It is common practice to try and keep the bit
weight just slightly more or less than the extending force. If
the extending force is far greater than the bit weight, or bit
weight far greater than extending force, the thrust bearings will
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fail prematurely. The life expectancy of the rotors and stators
is generally in the range of 1000 hours.
Turbine motors do not require a bypass or universal
joints. The rotation of the rotor shaft is concentric. Almost
all turbine bearing assemblies are mud lubricated.
Down-hole drilling motors have had limited success in
drilling operations due to a number of inherent problems:
1. unreliability of the thrust and radial bearing assembly;
2. limited load carrying capacities of the bearing assemblies;
3. low torque output of many motors,
. high rotational speed, causing short life of conical tri-cone
bits;
5. relatively poor high temperature capabilities;
6. poor bit cleaning, due to the low pressure drop restrictions
of many motor bearing assemblies, which limits the pressure
drop across the bit, hindering proper bit cleaning.
Premature motor failure or tri-cone bit failures neces-
sitate removal of the drill string from the well bore, costing
valuable rig time, and the replacement of the bit or motor, or
both, and the reinsertion of the dril] string into the well bore.
While a solution for conical tri-cone bits, namely to
slow the rotational speed, increase output torque and incorporate
a reliable robust bearing assembly, has been recognized, down-hole
driven motors have failed to meet expectations.
There are two types of radial and thrust bearings used
in down-hole motor bearing assemblies, ball or roller bearings,
and plain bearings. Plain bearings for radial applications are
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known as "journal bearings". Plain bearings used in down-hole
motors are either of a hard metal composition or consist of a
metal core covered with an elastomer. Ball and roller bearings
tend to erode in the drilling mud, and this erosion is propor-
tional to solids content in the mud.
Plain journal and thrust bearings tend to function
better operating in drilling mud than do roller or ball bearings,
but the torque that plain bearings oppose to rotation, because of
friction, is greater. This resisting torque is inversely propor-
tional to rotary speed. This makes plain bearings more applicable
to high speed turbine motors and roller or ball bearings more
applicable to slower speed positive displacement motors. The
friction of plain bearings is such that they do not lend their use
to slow speed motors. Roller and ball bearings do work with low
speed motors, but with a short life expectancy.
There is a requirement in the drilling industry for down
hole motors with a wide range of speeds to suit geological forma-
tions and 3 ma~or types of bits, tri-cone, diamond and polycry-
stalline.
Many down-hole motors utilize mud-lubricated bearing
assemblies that incorporate limited flow restrictors, and permit
approximately 10% of the drilling mud to flow through the bearing
assembly. However, such limited flow restrictors, permit only a
very limited pressure drop across them which, in turn, limits the
pressure drop across the drill bit, thereby hindering the proper
cleaning of the drill bit by the mud. Moreover, the drill mud
itself often contains sand or other abrasive solids which are
~257~
carried through the mud lubricated bearing assembly, thus causing
excessive wear and resulting in a relatively short life span for
the bearing assembly. As a result, the drill string must be
raised and lowered from the well bore for frequent repairs, thus
costing valuable rig time.
The problems inherent in the design of bearing assem-
blies for down-hole drilling motors are well described in
U.S. Patent No. 4,246,976, McDonald (Maurer Engineering, Houston),
issued on January 27, 1981.
Various methods have been proposed for sealing the
bearing assemblies of down-hole motors, so as to prevent this flow
of damaging drill fluid or mud. Prior art bearing assemblies have
often relied upon the exclusive use of elastomeric seals. Elasto-
meric seals, however, are known to result in premature wear, as
they are subjected to the abrasive drilling mud. As well, elasto-
meric sealing systems suffer from the further problem that the
pressure drop across the bit must be sealed by a seal within the
bearing system. As a result, the high pressure which is required
to clean the bit has been found to be excessive for the elasto-
meric seals, as it causes heat build-up and thereby results in
premature failure of the seals. Elastomeric seals are also sub-
ject to deterioration in high temperature applications, such a
geothermal wells, or where oil-based drilling muds are used.
The problem of heat build-up and the resultant premature
failure of seals for bearing assemblies is increased when either
a high-speed positive displacement motor or a turbine motor is em-
ployed in the drill string, both of which are highly desirable in
1257~365
the drilling industry. For example, turbine motors are particu-
larly useful in high temperature applications. However, as
turbines involve a high rate of output, this necessitates a speed
reducer or gear reduction assembly in order that it may be used
with other than diamond-tipped bits. Such gear reduction
assemblies also require an effective lubricating and sealing
arrangement.
One solution to the problem of bearing wear due to the
abrasive action of the drilling mud is to position the bearing
assembly within a non-abrasive lubricating fluid. In order to do
this, it is necessary to provide a special chamber for the bearing
assembly, sealed from the drilling mud. Sealed and lubricated
bearing assemblies are well known in the art. However, the means
for sealing a bearing chamber in this environment poses special
problems. Since the mandrel is moving rapidly in relation to the
casing, the seals must be dynamic. In addition, since down-hole
motors are often used under conditions of high temperature, and/or
in applications involving the use of hydrocarbon based or con-
taining drilling mud, the seals should not be subject to heat or
solvent induced deterioration.
There are also known designs for turbine motor gear
reduction assemblies that are provided within a leakproof oil
bath. As discussed in "Hydraulic Downhole Drilling Motors" by
W. Tiraspolsky, Gulf Publishing Company, at pp. 171-177, signi-
ficant research in this area is being conducted in the U.S.S.R.
One solution, as published by V/O "Licensintorg", 11, Ul. Min-
skaya, Moscow, 121108, U.S.S R., uses a sealed oil chamber with
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~257~6~
dynamic metal-to-metal face seals and a diaphragm to maintain the
chamber at the pressure of the mud in the string; see also Tiras-
polsky, Fig. 117, pp. 173 and 174 and U.S. Patent No~ 4,222,445.
The use of metal-to-metal dynamic seals is also shown,
for example, in U.S. Patent Nos. 3,971,450, 4,361,194, 4,372,400,
4,548,283, 4,593,774 and 4,577,704, as well as Canadian Patent
No. 1,144,202. Such rotating bearing seals for a down-hole motor
may be made of wear-resistant tungsten carbide.
It has also been recognized in the art that, in the use
of metal-to-metal dynamic seals, the pressure within the lubricant
chamber should be higher than that within the surrounding drilling
mud, whether within the drilling string or within the annulus.
(The mud pressure within the string is normally higher than that
within the annulus due to the pressure imposed by the pump. There
is a pressure drop across the bit due to the resistance to the
flow of mud between the bit and the bottom of the hole.) This is
so that any flow of fluid through the seals will constitute a flow
of lubricant into the mud, rather than the opposite. This wil~
keep the seal faces lubricated, and will avoid leakage of abrasive
drilling mud into the lubricant chamber.
One means used to pressurize the lubricant chambers
involves flexible membranes between the mud and the lubricant
material, as shown for example in U.S. Patents ~os. 4,593,774,
3,232,362 and 4,222,445, and as referred to above with reference
to the U.S.S.R. reduction unit. However, membrane seals are
readily subject to thermal and/or mechanical deterioration.
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Another solution is the use of a floating piston biased
against the lubricant chamber (e.g., U.S. Patent Nos. 4,361,194;
3,840,080; 4,090,574; 3,472,400; 4,246,976; 3,807,513; 3,971,~50;
4,493,381; and 4,577,704 and Canadian Patent No. 1,144,202). The
known biasing means include springs (U.S. Patent Nos. 4,361,194
and 4,090,574)-
A more sophisticated solution to the problem ofpressurization is shown in U.S. Patent No. 3,840,080, Berryman,
October ~, 1974, in which one face of the piston is exposed to the
mud within the string, normally subject to hydrostatic plus pump
pressure, and the other face is exposed partially to lubricant
pressure and partially to the mud pressure in the annulus (nor-
mally hydrostatic pressure only). Since the piston is balanced
this produces a pressure in the lubricant chamber higher than pump
plus hydrostatic (under normal operating conditions).
Nevertheless, further improvements in the sealing
ability and durability in metal-to-metal dynamic seals for
bearing/gear reduction chambers are still desirable, in order to
make the use of down-hole motors, especially high-speed displace-
ment motors and turbine motors, more practicable in a widervariety of applications.
A number of desirable criteria for a sealing system for
a lubricated bearing/gear reduction assembly for down-hole motors
may be identified:
(1) All dynamic seals should be mechanical, non-elastomeric
seals of wear-resistant material, for maximum durability and
sealing ability under a variety of conditions.
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~ 2 ~ 7 ~ ~ ~
(2) The pressure in the lubricant chamber should be higher
than that of the mud in the string or the annulus.
(3) Leakage of lubricant past the seals should be greater
than zero, in order to lubricate the seals, but minimal, in order
to minimize the need to replenish the lubricant fluid.
The present invention comprises a novel sealing system
designed to provide improved performance in meeting these cri-
teria.
The improvement or combination which is claimed as the
invention herein relates to a bearing assembly, or a bearing/gear
reduction assembly for a down-hole drill motor, the bearing
assembly including a relatively stationary casing and a rotary
driven mandrel disposed within it, and bearing means between the
mandrel and the casing, the bearing means being disposed in a
lubricant fluid within a lubricant chamber which is separated from
the drilling mud by means of one or more seal assemblies including
seal elements. In the improvement claimed, the lubricant fluid,
under normal operating conditions of the drill, is maintained at a
pressure higher than that of the drilling mud to which the sealing
areas of the seal elements are exposed, and in which all the
dynamic seal elements are made of non-elastomeric wear resistant
materials, and have sealing faces which are disposed in radially
extending planes. The invention also relates to such a sealing
system wherein the pressure within the lubricant chamber is main-
tained by means of a floating piston disposed between the mandrel
and the casing at one end of a lubricant chamber. The piston is
biased into the lubricant chamber by a biasing force and is
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12578~5
divided into two parts, one part being rotationally stationary
with reference to the casing and one part being rotationally
stationary with reference to the mandrel, and a non-elastomeric
wear resisting dynamic seal assembly is provided between the two
parts of the piston to seal the lubricant chamber from the
drilling mud. Other features of particular embodiments of the
invention are defined in the claims appended to this specifica-
tion, which set forth the exclusive rights claimed by the
applicant.
The invention will now be described by way of example
only, reference being had to the accompanying drawings in which:
Figure 1 is a longitudinal sectional view of an embod-
iment of this invention, including a bearing assembly and a gear
reduction assembly;
Figure 2 is a longitudinal sectional view of an alter-
nate piston construction in a bearing-only tool;
Figure 3 is a longitudinal sectional view of an alter-
nate gear reduction arrangement;
Figure 4 is a cross-sectional view, taken along lines 4-
4 of Flgure 1, of the clutch assembly of the piston of the present
nventlon;
Figure 5 is a cross-sectional view, taken along lines 5-
5 of Figure 1, of the gear reduction assembly of one embodiment of
the present invention;
Figure 6 is a sectional view showing an alternate seal
assembly at the lower end of the construction shown in Figure 1;
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~257~36~
Figure 7 is a sectional view showing another alternate
seal assembly at the lower end of the construction of Figure 1.
Referring to the accompanying drawings, the basic struc-
ture of the seal means and the bearing assembly is shown in
Figure 1, wherein an output shaft 1, which is driven by a down-
hole positive displacement motor (not shown), extends to a univer-
sal joint assembly 2. (In the case of a turbine motor, the output
shaft is concentric, and no universal joint is needed~ The
presence or absence o~ a universal joint is not material to the
invention described herein.) The output shaft 3 of the universal
joint assembly 2 is threaded to the driven mandrel 4 within the
housing 5 of the drill stringO (Numeral 5 designates the entire
casing of the bearing/gear reduction assembly. Casing 5 is com-
posed of a number of separate components.) Drilling fluid or mud
is pumped down the drill string at a high pressure through the
motor and thereafter flows down a chamber 6 between universal
joint 2, output shaft 3 and the mandrel 4 and the outer casing 5.
A passageway 7 connected to the chamber 6, permits the mud to flow
into the center bore of the drill string, bypassing the bearing/
gear reduction chamber 8. The mud continues down the center
bore 9 of the drill string to the drill bit (not shown), which is
secured to the lower end 10 of the driven mandrel 4. (Numeral 4
designates the entire driven mandrel within the bearing/gear re-
duction assembly. Mandrel 4 is comprised of a number of separate
components.) The mud flows outwardly across the drill bit, there-
by facilitating both the drilling action and the removal of
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debris. It thereafter rises up the annulus around the drill
string to the surface for treatment and recirculation.
Extending longitudinally within the drill string is a
chamber 8 created between the substantially stationary outer
casing 5 of the drill string and the driven inner mandrel 4. At
the upper end of this chamber is a dynamic, mechanical seal
assembly 11, which shall later be described in greater detail.
This seal assembly is incorporated in a floating piston 12, which
is exposed at its upper end to the drilling mud which flows down
the cavity 6 of the drill string. At the lower end of the
chamber 8 is another dynamic, mechanical seal assembly 13. The
lower end of the seal assembly 13 is exposed to the drilling mud
that flows upwardly through the well bore, after having passed
down the drill string and across the drill bit. The chamber 8,
which is sealed at either end by the upper and lower dynamic seal
assemblies 12 and 13, is filled with a lubricating fluid, by means
of a fill hole 14 through the exterior of the casing of the drill
string. A viewing port 15 is provided in order to view the posi-
tion of the floating piston 12 and hence determine whether or not
chamber 8 is filled with lubricant to the optimum level. As shown
in Figure 1, the level is optimal. The lubricating fluid may be
oil or any other lubricant adapted for use in down-hole drilling
tools. Such lubricants are well known in the art.
Positioned within the lubricant chamber 18 are a series
of bearing assemblies 15, 16, 17, 18 and 19 composed of radial and
axial bearings. There is also provided within the lubricant
chamber 8 a planetary gear reduction assembly 20. A gear reduc
~2S7~36~
tion assembly i5 desirable when the down-hole motor is either a
high-speed positive displacement motor or a turbine motor, as it
slows the output rotary speed o~ the bearing assembly while in-
creasing the output torque. In the Figure 1 embodiment, there is
a 3:1 planetary gear reduction assembly that slows the output
speed to 1/3 the input speed, while tripling the output torque.
In the gear reduction area, an output shaft is provided
by an extension 21 of the tube 22 between the mandrel 4 and the
outer casing 5. Rotatably mounted on the tube 22 are the planet-
ary gears 23 that engaqe splines formed on the casing 5 and themandrel ~ and are thereby driven to rotate between the stationary
casing 5 and the rotating mandrel 4. The tube 22 drives a second
section 24 of the mandrel 4. Another dynamic, mechanical seal
assembly 25 is required between the two mandrel sections, as a
result of the 3:1 speed reduction. The gear reduction assembly
and its relationship to the casing 5 and the mandrel 4 is illu-
strated in cross-section in Figure 5. In Figure 1, the gear
reduction assembly is foreshortened to save space. In the pre-
ferred embodiment hereof, three sets of planetary gears are
2 provided for greater strength in transmitting the increased torque
to the output tube 22.
It is important to the sealing of the chamber 8, in
which the bearing and gear reduction assemblies are provided, that
the lubricating fluid contained therein is at a higher pressure
than that of the drilling mud, both inside the drill string and
flowing externally in the well bore. By maintaining the lubri-
cating fluid within the chamber at a higher relative pressure, it
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~25786S
70242-6
is apparent that this fluid will tend to flow externally from the
chamber, rather than there being an ingress of drilling mud into
the chamber.
The upper seal assembly 11 is provided within a pressure
balanced and slidable piston 12. The piston 12 is exposed at its
upper end over area Al to a pressure that is equal to the com-
bination of the hydrostatic pressure of the mud plus the pump
pressure. At its lower face, the piston 12 is exposed to a com-
bination of hydrostatic mud pressure on a small area A2 and
lubricant pressure on the remaining area A3. The areas are such
that Al = A2 + A3. Hence,
AlPH+p = A2PH + A3PL
where
PH+p = hydrostatic plus pump pressure
PH = hydrostatic pressure
PL = lubricant pressure.
As the hydrostatic pressure on the small area A2 f the face is
less than the combined hydrostatic plus pump pressure on the upper
end of the piston, the pressure on the oil within the chamber is
greater than the combined pressure of the hydrostatic plus pump
pressures of the mud in the string. Accordingly, there tends to
be a slight leakage of oil from the chamber through the dynamic
seals and into the mud.
It will be apparent that the dynamic pressure of the mud
flow down the annulus 6, impacting on area Al of the piston, will
of itself tend to bias the piston into the lubricant chamber. The
pressure exerted by the mud flow can be varied by varying the
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70242-6
design of ports 7. Conventionally, such ports are large and are
angled toward the direction of the flow, to make it as easy as
possible to divert the mud into the interior of the string. This
design obviously minimizes the impact of the flow on the piston.
The right-angled ports 7 as shown in Figure 1 create a higher
impact pressure. Small ports angled away Erom the flow would
increase the pressure further.
If the stepped structure of the piston is not used, the
bore of the barrel can be uniform, and the piston can be short-
ened, as shown in Figure 2, where, however, a biasing spring isused.
The lower portion 26 of the seal assembly 11 is locked
onto the mandrel by a clutch mechanism 27 which is activated when
the mandrel rotates. As shown in Figure 4, the clutch mechan-
ism 27 comprises a roller bearing 28 that is provided within a
narrowing groove 29 so that when the mandrel 4 is driven in the
proper drilling direction, the roller bearing 28 will tend to roll
into the narrow end of the groove 29, thereby frictionally
.
. ,,~ .
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~257F~65
70242-6
causing the lower portion 26 of the seal assembly 11 to rotate
with the mandrel 4.
The piston 12 is non-rotatably engaged with the station-
ary casing 5 of the drill string by contact friction only, which
has been found to be sufficient to preclude relative rotational
movement between the piston and the casing. The piston could
obviously also be positively locked into the casing by splines or
other such conventional devices, or by a clutch mechanism like
mechanism 27. The upper part 30 of the seal assembly 11 is in
turn non-rotatably locked into the piston by suitable means,
which, as shown in Figure 1, may be a spline and key arrange-
ment 31. In order to maximize wear resistance, dynamic rotary
face seal elements 32 and 33 are provided within the seal
assembly 11. The dynamic face seal elements 32 and 33 are made
from a non-elastomeric, wear resistant material such as tungsten
carbide. They are lubricated by the small amount of lubricant
that tends to flow from the high pressure, lubricant-filled
chamber 8. The faces of the seal elements 32 and 33 extend radi-
ally. The tungsten carbide face seal elements 32 and 33 are
backed by steel back-up rings 34 and 35 to which they are non-
rotatably connected. Seal element 32 is fixed to back-up ring 34
by pins 36, whereby it is axially movable relative to the back-up
ring 34. It is urged against seal element 33 by springs (not
shown) interposed between element 32 and its back-up ring 34. An
axial bearing 37 is provided between the portion of the seal
assembly that rotates with the mandrel and the portion that is
non-rotatably engaged with the casing. A radial bearing 38 is
æ~
,.,, ~ ~
- 18 -
~25786~
70242-6
provided between the piston 12 and the mandrel 4. Various sta-
tionary elastomeric seals are provided as shown.
By dividing the piston into two parts, one of which
rotates with the mandrel, and one oE which is non-rotatably en-
gaged with the casing, this invention avoids the need for any
rotary elastomeric seals on the piston. All the seals on the
piston except the seal between elements 32 and 33 are static
elastomeric seal.s. In some prior art devices, the use of dynamic
elastomeric seals on a floating piston is avoided by placing the
piston in a separate annular chamber formed by an additional tube
non-rotatably secured within the casing. This complication is not
necessary in the present construction.
In the gear reduction area, the seal assembly 25 that is
provided between the separate mandrel sections 24 and 39 that are
rotating at different relative speeds also include dynamic seal
elements of a non-elastomeric wear resistant material such as
tungsten carbide. The seal assembly 25 is virtually identical in
the general principles of its construction to the upper seal
assembly 11. The seal elements 40 and 41 are coupled by pins to a
back-up ring 42 and mandrel section 24 respectively. Springs (not
shown) are provided between element 40 and back-up ring 42 to bias
the rotary face seal elements 40 and 41 against one another. A
key and spline arrangement 43 is provided to couple seal ele-
ment 40 to -the mandrel section 39. A port 44 is provided to allow
lubricant to reach the seal assembly.
The dynamic seal assembly 25 is not necessary when a
gear reduction unit is not employed, i.e., when the present inven-
~A
-- 19 --
~257865
tion is practised with a bearing assembly only. The entire gear
reduction assembly and seal assembly 25 can be removed without
departing from the practise of this invention.
Provided towards the bottom end of the drill string in
the embodiment of Figure 1 is another dynamic seal assembly 13.
As with the other seal assemblies that are provided within the
drill string, the seal assembly 13 is provided with rotary face
seal elements 45 and 46, of a non-elastomeric wear resistant
material such as tungsten carbide, that are lubricated by the
limited flow of the higher pressure lubricating oil from the
lubricant chamber to the interior of the drill string. Rotating
seal element 45 is non-rotatably pinned to a back-up ring 47 by
pins 48, with springs (not shown) being provided between ele-
ment 45 and ring 47. Ring 47 is affixed to the driven mandrel by
means of a set screw 49. Biased against this rotating seal ele-
ment 45 is the stationary seal element 46 that is non-rotatably
pinned to a back-up ring 50 affixed to the stationary casing of
the drill string by means of a key and spline arrangement, similar
to those used on other seal assemblies herein.
In the Figure 1 embodiment, the bottom seal assembly 13
is exposed on one side to the lubricant in the chamber 8 and on
the other side to the drilling mud in the interior chamber 9 of
the string. This is accomplished by means of an internal
passage 51 from the center bore of the drill string to one face of
the dynamic seal assembly, in which is placed a flow rate limiting
device 52. The primary face seal between elements 45 and 46 is
sealed from mud in the annulus by stationary elastomeric seals 53
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~:25i~8~i
and 54 and by a second dynamic seal formed between element 50 and
secondary element 55. Element 55 is a non-elastomeric wear
resistant element formed, for example, of tungsten carbide, and
non-rotatably fixed to element 45. The pressure of the mud in the
string, as with the pressure of the mud external to the upper seal
assembly, is equal to the sum of the hydrostatic plus pump
pressures. Therefore, as the pressure of the oil is constant
throughout the lubricant chamber 8, it is of the same relatively
higher pressure as the lubricant within the upper end of this
chamber. Accordingly, the lubricant will tend to flow externally
through the lower seal assembly from the lubricant chamber to the
interior of the string, rather than there being an ingress of
drilling mud. However, the pressure drop is less than if the seal
were exposed to the mud in the annulus.
The advantage of the Figure 1 arrangement of the lower
seal assembly is that the rate of flow is somewhat less than it
would be if the seal were exposed on one side to the mud in the
annulus, which would result in a higher pressure drop across the
seal and therefore a higher rate of lubricant flow across the seal
face.
Figure 2 shows a drill string bearing only assembly with
an alternate floating piston design. The lower seal arrangement
is not shown in Figure 2, but may consist of any of the arrange-
ments described herein. In the Figure 2 embodiment, the floating
piston 12 has the same areas on both its upper and lower faces and
is exposed on its lower face only to the lubricant chamber 8. It
is biased into the lubricant chamber 8 by a helical spring 56,
~57865
whereby the pressure in chamber 8 exceeds that in chamber 6 so
that lubricant tends to flow from chamber 8 into chamber 6, and
mud does not tend to flow into chamher 8. Spring 56 is biased at
its upper end against casing 5 and is held in place by ring 57.
As lubricant leaks out of chamber 8, the piston 12 moves into the
chamber with spring 56 elongating as it follows the piston. In
any spring, force decreases as elongation increases. Therefore,
in the Figure 2 embodiment, the pressure differential between
chambers 8 and 5 decreases as chamber 8 empties. This is a
disadvantage which is avoided by the Figure 1 construc-tion.
In Figure 2, the gear reduction assembly 20 and second
dynamic seal assembly 25 are eliminated, and only a bearing func-
tion is provided. The individual elements of the bearing assembly
of Figure 2 will be readily understood by persons skilled and need
not be described in detail herein. The section of the bearing
assembly indicated in Figure 2 at 58 is described in detail and
claimed in applicant's co-pending Canadian application
No.
The structure shown in Figure 3 is the same as that
shown in Figure 1, except that an additional gear reduction
stage 59 is provided. The second planetary gear stage 59 effects
a total 9:1 gear reduction, which is especially well suited to
turbine motors, as output speed is greatly decreased and output
torque is greatly increased. In order to achieve the 9:] gear
reduction, a single 3:1 planetary gear assembly 59 is provided in
addition to the triple planetary gear assembly 20 of Figure 1. In
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~257~65
this embodiment, there is also provided a fourth mechanical seal
assembly 60, which is made necessary by the second planetary gear
assembly. Assembly 60 is the same in structure as the third seal
assembly 25, previously described. In this embodiment, the upper
gear reduction assembly 59 requires only a single set of planetary
gears because it is designed for high speed, low torque input and
output loads. The lower assembly 25 is designed for high output
torque and is therefore preferably a three set assembly.
In the Figure 6 embodiment of the lower seal assem-
bly 13, the passage 51 for the drilling mud from the center bore 9
of the drill string to the seal assembly 13, as shown in Figure 1,
is eliminated. Rather, the drilling mud to which is exposed one
face of the seal assembly is external to the drill string, flowing
upwardly through the external annulus 61 of the well bore. The
same principles of the differential pressure between the oil with-
in the chamber and the external mud apply, as in the bottom seal
arrangement 13 discussed with respect to Figure 1. However, the
pressure drop across the seal is normally higher in this embod-
iment than in the embodiment of Figure 1, since annulus pressure
is normally lower than string interior pressure. The detailed
structure of the sealing assembly 13 of Figure 6 is the same as
that of Figure 1, except that the special structures 51, 52
and 55, adapted to expose the Figure 1 assembly to string mud, are
eliminated.
Figure 7 illustrates as alternate bottom mechanical seal
assembly 13. In this embodiment, an additional stationary elasto-
meric seal 62 seals the interface between the seal element 46 and
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~57~
its backing ring 50 from the drilling mud, allowing lubricant to
penetrate into this area. This provides a means to reduce distor-
tion of the seal element 46 as a result of being pressed against
the backing ring 50. It is apparent that by providing this fea-
ture, seal element 46 is slidable or floating on its backing ring,
thereby reducing the direct mechanical force between the element
and the backing ring. Rather, the force on the backing ring in
this embodiment is substantially hydraulic in nature. This re-
duces the tendency for the very precisely machined seal elements
to be forced against the less precisely machined backing rings.
This in turn is believed to reduce deformation of the seal ele-
ments and thereby reduce oil leakage and increase seal life and
prolong the intervals between servicing of the bearing/gear
reduction assembly.
The above description covers the preferred forms of the
invention. Many modifications of the principles of this invention
are possible within its scope.
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