U.S. patent application number 10/756288 was filed with the patent office on 2005-07-14 for method and apparatus for preventing the friction induced rotation of non-rotating stabilizers.
This patent application is currently assigned to VALIDUS. Invention is credited to Schuh, Frank J..
Application Number | 20050150694 10/756288 |
Document ID | / |
Family ID | 34739802 |
Filed Date | 2005-07-14 |
United States Patent
Application |
20050150694 |
Kind Code |
A1 |
Schuh, Frank J. |
July 14, 2005 |
Method and apparatus for preventing the friction induced rotation
of non-rotating stabilizers
Abstract
A method and apparatus prevents rotation of a stabilizer which
rotates due to friction-induced forces of bearing seals disposed in
a shaft. The apparatus includes a stabilizer having a ridged edge
where the stabilizer is mounted to a drilling tool using a
deformable member. The deformable member assists in creating
friction of the stabilizer edge against a drilled bore hole wall
and thereby reduces friction-induced rotation.
Inventors: |
Schuh, Frank J.; (Plano,
TX) |
Correspondence
Address: |
SUGHRUE MION, PLLC
2100 PENNSYLVANIA AVENUE, N.W.
SUITE 800
WASHINGTON
DC
20037
US
|
Assignee: |
VALIDUS
|
Family ID: |
34739802 |
Appl. No.: |
10/756288 |
Filed: |
January 14, 2004 |
Current U.S.
Class: |
175/325.1 ;
175/325.3; 175/325.5 |
Current CPC
Class: |
E21B 17/1057
20130101 |
Class at
Publication: |
175/325.1 ;
175/325.3; 175/325.5 |
International
Class: |
E21B 017/10 |
Claims
I claim:
1. A stabilizer assembly mountable on a surface of a drilling tool
having a longitudinal shaft, said stabilizer comprising: a
plurality of stabilizer blades, each blade having multiple ridges
and being independently mounted on the drilling tool.
2. The stabilizer assembly according to claim 1, further comprising
a displacement member disposed between each stabilizer blade and
the drilling tool, said displacement member deformable to provide a
minimum radius and a maximum radius, different from the minimum
radius, from a center of the longitudinal shaft to an outer
extremity of at least one of the multiple ridges.
3. The stabilizer assembly according to claim 2, wherein when the
displacement member provides the minimum radius, at least one of
the multiple ridges are exposed.
4. The stabilizer assembly according to claim 2, wherein the
displacement member comprises a load spring.
5. The stabilizer assembly according to claim 4, wherein the load
spring provides a rotational resistance RR sufficient to
counterbalance frictional rotation RDT according to an expression:
4 RDT = bd ff SBF 2 + 2 st where:
9 RDT = rotational driving torque in lbs. bd = shaft bearing
diameter in. ff = bearing friction factor * SBF = sum of all the
bearing forces lbs. st = rotating seal torque in lbs.
6. The stabilizer assembly according to claim 4, wherein the load
spring has a spring load in a range of 50-60 lbs across a travel
distance of 1/8 inch.
7. A stabilizer assembly mountable on a surface of a drilling tool
having a longitudinal shaft, said stabilizer comprising: at least a
first stabilizer blade having multiple ridges and movably mounted
on the drilling tool.
8. The stabilizer assembly according to claim 7, wherein the
stabilizer assembly includes a second stabilizer blade having
multiple ridges, said second stabilizer blade independently and
movably mounted onto the drilling tool from the first stabilizer
blade.
9. The stabilizer assembly according to claim 8, wherein each of
the first and second stabilizer blades is attached to the drilling
tool by a displacement member disposed between the stabilizer blade
and the drilling tool, said displacement member deformable to
provide a minimum radius and a maximum radius from a center of the
longitudinal shaft to an outer extremity of at least one of the
multiple ridges.
10. The stabilizer assembly according to claim 9, wherein when the
displacement member provides the minimum radius, at least one of
the multiple ridges are exposed.
11. The stabilizer assembly according to claim 9, wherein the
displacement member comprises a load spring.
12. The stabilizer according to claim 11, wherein the load spring
provides a rotational resistance RR sufficient to counterbalance
frictional rotation FR according to an expression: This equation
must be replaced by equation 3 5 RT s = sd 2 ( FS + F spring ) (
0.63 ) ( 0.9 ) where:
10 RTs = rotation resisting torque with spring loaded contacts in
lbs. FS = total lateral loads on all of the stabilizer blades lbs.
sd = stabilizer diameter in. .SIGMA.Fspring = sum of all of the
spring forces lbs
13. The stabilizer according to claim 11, wherein the load spring
has a spring load in a range of 50-60 lbs across a travel distance
of 1/8 inch.
14. A method of forming a stabilizer assembly for a drilling tool
including a plural number of bearings disposed in a longitudinal
shaft, said method comprising: determining a lateral force of each
plural number of bearings SBF; determining a shaft bearing diameter
bd; determining a frictional seal torque st; determining a bearing
friction factor ff; and determining a resistance torque RTs for the
stabilizer sufficient to counterbalance a frictional rotation RDT
according to an expression: 6 RT s = sd 2 ( FS + F spring ) ( 0.63
) ( 0.9 ) where:
11 RTs = rotation resisting torque with spring loaded contacts in
lbs. FS = total lateral loads on all of the stabilizer blades lbs.
sd = stabilizer diameter in. .SIGMA.Fspring = sum of all of the
spring forces lbs.
Description
FIELD OF THE INVENTION
[0001] This invention provides a novel method and apparatus for
preventing the friction induced rotation of non-rotating
stabilizers which are attached to a drilling tool to control the
direction and orientation of drilling. The invention can be
utilized to prevent rotation at any combination of hole angle,
curvature rate and bit load. The present invention provides a more
efficient method of operating rotary steerable directional
tools.
BACKGROUND
[0002] Most rotary steerable systems utilize non-rotating
stabilizers to control the trajectory of the hole. The rotating
friction between the non-rotating stabilizer and the shaft that
turns the bit on conventional systems causes the non-rotating
stabilizer to rotate in a clockwise direction. With conventionally
surfaced stabilizers, the procession rate is related to the ratio
of the rotational friction force between the shaft and the fixed
stabilizer to the axial drag force between the fixed stabilizer and
the borehole wall. The frictional rotation rate decreases as the
hole angle, curvature rate, and/or bit weight increases. However,
rotation rates may become excessive at low hole angles, low
curvature rates and/or low bit weights. The worst conditions are
most likely to occur at the kick off point in a vertical hole. This
problem prevents the use of conventional rotary steerable systems
on many directional drilling applications.
[0003] For example, Table 1 shows the expected frictional rotation
rates for a 12 ft non-rotating stabilizer that includes a
conventional smooth surfaced adjustable stabilizer, a fixed
stabilizer and which utilizes low friction sealed bearings between
the shaft and the non-rotating unit.
1TABLE 1 CONVENTIONAL NON-ROTATING STABILIZER Hole Curvature Bit
Lateral/axial Frictional Angle Rate Weight Slide friction Rotation
rate deg. deg/100 ft kips Ratio deg/axial ft 30.0 6 25 1 14.5 30.0
3 25 1 17.1 30.0 2 25 1 18.4 0.5 3 25 1 37.4 0.5 2 25 1 48.1 0.5 3
10 1 55.5 0.5 2 10 1 68.6
[0004] As suggested by the last column, the adjustable stabilizer
blades must be continuously adjusted to compensate for the
friction-induced rotation of the non-rotating unit. Even with a
hole angle of 30 degrees the expected frictional rotation rates
would be a problem. With a 0.5 degree hole angle, the rotation
rates are unacceptable.
SUMMARY OF THE INVENTION
[0005] Applicant solves the frictional rotation problem by making
the stabilizer surface act like a drag bit and by increasing the
contact forces between the stabilizer and the bore wall formation.
Rotation is prevented whenever the threshold torque required to
rotate the drag bit like contacts exceed the rotational driving
torque. The design is so effective that it can prevent frictional
rotation by only applying the design concept to the fixed
stabilizer.
DESCRIPTION OF THE DRAWINGS
[0006] Preferred embodiments of the invention will be described
below in reference to the appended drawings wherein,
[0007] FIGS. 1A-1D illustrate the relationship between rotational
drive torque, bearing forces and frictional rotation rate as
observed by the present inventor;
[0008] FIG. 2 illustrates a fixed stabilizer as mounted on a
drilling tool according to a preferred embodiment;
[0009] FIGS. 3A-B illustrate the fixed stabilizer according to a
preferred embodiment;
[0010] FIGS. 4A-B illustrate the forces on a PDC cutter which
models drag behavior of the fixed stabilizer according to the
present invention.
DESCRIPTION OF PREFERRED EMBODIMENT
[0011] Referring to the accompanying Figures, a preferred
embodiment of the invention is described as follows.
[0012] Referring to FIGS. 1A-1D, the present inventor determined
that the mechanics of frictional rotation are defined generally by
the following equations: 1 RDT = bd ff SBF 2 + 2 st ( 1 )
[0013] Where:
2 RDT = Rotational driving torque in lbs. bd = Shaft bearing
diameter in. ff = Bearing friction factor * SBF = Sum of all the
bearing forces lbs. st = Rotating seal torque in lbs
[0014] The present invention is specifically applicable to drilling
tools including two to four bearings. However, the invention may be
applied to a drilling tool with a different number of bearings as
long as the summation of the lateral forces on the bearings (SBF)
is taken into account. The rotational driving torque comes from the
frictional torque in the Kalsi seals and the lateral contact forces
in the bearings between the shaft and the (non-adjustable
stabilizer) (NAS). The resisting forces are generated by the
lateral contact forces between the stabilizers and the hole.
Whenever the resisting forces are larger than the frictional forces
rotation is prevented.
[0015] FIG. 2 illustrates the placement of fixed stabilizer blades
1, 1! on a non-rotating stabilizer according to the present
invention. Reference number 2 corresponds to an adjustable
stabilizer to position the drilling tool in the bore hole.
[0016] The present inventor noted that the calculated cases for
Table 1 (conventionally surfaced stabilizer) assumed that the axial
sliding friction factor and the rotating friction factor were
equal. If the sliding surface of the stabilizer blade were modified
to increase the rotating friction factor, the rotation rate would
be reduced.
[0017] Table 2 shows the effect of utilizing a blade surface that
provides a rotational friction factor that is 3 times the axial
sliding friction factor. The desired effect is enhanced by aligning
the edges of the ridges parallel to the axis of the bore hole and
making the ridges sharp.
3TABLE 2 NON-ROTATING STABILIZER WITH DIRECTIONAL FRICTION Hole
Curvature Bit Lateral/axial Frictional Angle Rate Weight Slide
friction Rotation rate deg. deg/100 ft Kips Ratio deg/axial ft 30.0
6 25 3 4.8 30.0 3 25 3 5.7 30.0 2 25 3 6.1 0.5 3 25 3 12.5 0.5 2 25
3 16.0 0.5 3 10 3 18.5 0.5 2 10 3 22.9
[0018] This improvement makes the frictional rotational rates
acceptable in 30 degree holes but still presents a significant
problem in 0.5 degree holes, especially at reduced bit weight and
curvature rate.
[0019] In the present invention, the contact surface of the
stabilizer blade is modified to inhibit lateral movement. The
preferred modification places axial ridges on the surface of fixed
stabilizer blades. The lateral forces on the stabilizer push the
ridges into the bore wall, thereby preventing lateral rotation of
the drilling assembly whenever the resisting shear forces in the
formation wall exceed the rotational friction force.
[0020] Referring to FIG. 3A, each stabilizer fin has 6 sharp drag
bit shaped cutters 3a-3f. The cutters are equally spaced from the
tool center. The cutters are curved along the axial direction.
Under low loads only a single cutter will contact the wall of the
hole. Referring to FIG. 3B, the stabilizer fins are supported on
load springs. The allowable radial travel is set to provide an
under gauge diameter (relative to the bore hole) when the blades
are fully collapsed and an over gauge diameter when fully extended.
The trailing edge will be 30.degree. (angle A) below the tangential
surface.
[0021] The cutters act like polycrystalline diamond compact (PDC)
cutters on a PDC bit. The rotational mechanics of this design can
be modeled using technology developed for the drill bit industry.
FIGS. 4A-4B illustrate known configurations for PDC cutters. An
excellent source of useful information was published by Glowka of
Sandia Nat'l Labs in the Society of Petroleum Engineers Journal of
Petroleum Technology in August 1989 pgs 797-849.
[0022] Glowka used a variety of single PDC cutters to measure the
mechanics of drilling in three kinds of rock. The test included
flat faced cutters as well as sharp edge cutters. Most of the tests
measured the axial cutter loads and the penetration forces as a
function of the depth of cut.
[0023] They developed the following empirical relationships for
cutting dry rock at the surface.
[0024] FDB=FA.multidot.(0.90+2.2.multidot.D)
[0025] FDT=FA.multidot.(0.65-0.58.multidot.D)
[0026] FDS=FA.multidot.(0.63+0.88.multidot.D)
[0027] where
4 FDB = Cutter drag force in Berea Sandstone lbs. FDT = Cutter drag
force in Tennessee Marble lbs. FDS = Cutter drag force in Sierra
White Granite lbs. FA = Downward force on the dull cutter lbs. D =
Depth of cut in.
[0028] The tests that used sharp cutters required larger cutter
drag forces than observed with the dull cutter tests. They also ran
tests with drilling fluid. These tests showed that the drilling
fluid acted as a lubricant and reduced the cutter drag forces by 10
percent.
[0029] The inventor notes that the cutter drag forces are greater
in softer rocks. Using the performance in granite should
underestimate the cutter drag forces in all oilfield formations.
Combining all these factors gives the following safe estimate for
the rotational resistance of the stabilizer design of the
invention: 2 RT = FS ( 0.63 ) ( 0.9 ) sd 2 ( 2 )
[0030] Where:
5 RT = Rotation resisting torque in lbs. FS = Total lateral loads
on all of the stabilizer blades lbs. sd = stabilizer diameter
in.
[0031] By using a bit cutter like contact, the mechanics are
changed from a two dimensional sliding problem to establishing a
threshold resisting load that prevents any rotation whenever it
exceeds the driving torque RDT defined above in equation (1).
[0032] Referring back to FIG. 3B, the fixed stabilizer adjusts from
8 3/8 in. outer diameter to 8 5/8 in. outer diameter for drilling 8
1/2 in. holes. The fixed stabilizer load springs apply 50 to 60
pound loads across the travel limits. The shaft uses three low
friction bearings. Both the cutter like contacts and the spring
assisted fixed stabilizer blades are needed to completely eliminate
frictional rotation. The rotational torque in this situation is
modified from equation (2) above to further include the sum of the
spring forces. 3 RT s = sd 2 ( FS + F spring ) ( 0.63 ) ( 0.9 ) . (
3 )
[0033] Where:
6 RT.sub.s = Rotation resisting torque with spring loaded contacts
in lbs. FS = Total lateral loads on all of the stabilizer blades
lbs. sd = stabilizer diameter in. .SIGMA.Fspring = Sum of all of
the spring forces lbs.
[0034] Table 3 shows the expected performance of using cutter type
contacts without spring loaded stabilizer blades.
7TABLE 3 NON-ROTATING STABILIZER WITH ROTATION AVOIDANCE CUTTERS ON
THE FIXED STABILIZER Rotational Rotational Avoidance Hole Curvature
Bit Driving Resisting Design Angle Rate Weight Torque Torque Design
deg. deg/100 ft kips In lbs In lbs Factor 30.0 6 25 145 1349 9.3
30.0 3 25 131 902 6.9 30.0 2 25 126 750 5.9 0.5 3 25 130 375 2.9
0.5 2 25 126 227 1.8 0.5 1 25 120 73 0.6 0.5 3 10 128 27 0.2 0.5 2
10 124 23 0.2 0.5 1 10 120 14 0.1 0.5 3 5 129 110 0.9 0.5 2 5 125
73 0.6 0.5 1 5 120 37 0.3
[0035] This design easily prevents rotation at both 30 degree and
0.5 degree hole with bit weights of 25,000 lbs and curvature rates
of 2 deg/100 ft or more. However, the last five cases in the table
would not stop frictional rotation. As shown in Table 4, adding 50
to 60 lb. springs to each of the stabilizer blades completely
eliminates any chance of frictional rotation. Five blades are
contemplated for the preferred embodiment. At a minimum, the
present invention can minimize rotation to 1-3.degree. of roation
per foot drilled, even for a verticle hole.
8TABLE 4 NON-ROTATING STABILIZER WITH ROTATION AVOIDANCE CUTTERS
AND 50/60 POUND SPRINGS ON THE FIVE FIXED STABILIZER BLADES
Rotational Rotational Avoidance Hole Curvature Bit Driving
Resisting Design Angle Rate Weight Torque Torque Design deg.
deg/100 ft kips In lbs In lbs Factor 30.0 6 25 145 2020 14.0 30.0 3
25 131 1567 11.9 30.0 2 25 126 1413 11.2 0.5 3 25 130 1040 8.0 0.5
2 25 126 891 7.1 0.5 1 25 120 736 6.1 0.5 3 10 128 699 5.5 0.5 2 10
124 691 5.6 0.5 1 10 120 679 5.7 0.5 3 5 129 779 6.0 0.5 2 5 125
740 5.9 0.5 1 5 120 700 5.8
[0036] The combination of cutter like contacts and spring loaded
blades provides a rotational resistance force that is at least 5
times greater than the frictional driving force under all
conditions.
[0037] While a preferred embodiment has been described above, one
skilled in the art would recognize that the invention can be
modified and still fall within the scope of the appended claims.
For instance, the load spring can be replaced by alternative
mechanism to exert a lateral force against the wall such as a
hydraulic system.
* * * * *