JPS5914159B2 - Vertical drilling equipment - Google Patents

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to a vertical drilling motor, such as a turbo drill or motor driven drill.

Vertical drilling motors were invented 100 years ago and were widely used in the 1920's.

Vertical drilling motors did not find widespread use until the 1950s, when turbo drills began to be used in the Soviet Union.

It is estimated that 85 wells in the Soviet Union were drilled with turbo drills in the early 1960s.

Vertical drilling motors were widely used in the United States for straight drilling, but were rarely used for vertical drilling due to bearing and sealing deficiencies.

Commercial vertical drilling motors have roller bits of 50 to 1
It works most efficiently at 50rpm, but from 300rpm
It operates in a speed range of 11000 rpm.

Since the motor operates at high speed, the roller bearing, which can be used for 200 hours at low speed, will break in 5 to 15 hours.

Vertical drilling motors have been effectively restricted from use in oil wells and other applications due to problems with radial and hysterus bearings, lubrication systems, turbine efficiency, jacket construction, etc.

Vertical drilling motors were invented shortly after the advent of rotary drilling equipment in the 1960's.

A very primitive turbo drill is shown in US Pat. No. 174,922.

U.S. Pat. No. 2,928,888 shows a single stage axial flow turbodrill that is similar in some respects to current turbodrills.

U.S. Pat. No. 1,482,702 shows a multi-stage turbo drill that was a precursor to the turbo drills currently in use.

The turbo drill includes a lubrication system for operating the thrust bearing in oil or grease.

The turbulent fluid acting on the floating piston pressurizes the lubricating oil in the system.

Current turbo drill bearings operate in the drilling mud, which causes them to break in a short period of time, limiting their range of applications.

US Pat. No. 1,681,094 shows a single gear turbo drill.

This turbo drill was used in the Soviet Union from 1924 to 1995.
It was widely used for 34 years.

Although there were many problems, multistage turbo drills were perfected in the 1940s and 1950s, and 80 to 90 wells in the Soviet Union were drilled using axial flow turbo drills.

Licenses for turbodrill technology were granted to the United States, France, Germany, and Austria.

Turbo drills were not yet fully accepted commercially and were mainly used for straight drilling.

All vertical drilling motors essentially consist of four basic parts:

That is, (1) motor part (2) Thrust bearing part (3) Radial bearing part (4) Rotating seal part It is.

The bearings and seals can be placed as a separate unit in the motor section and can be used with any type of motor.

Motors are divided into the following two types.

(1) Turbo drill (2) Positive motor (PO8ITIVE DISPLAC
A turbo drill uses the moment change of the turbulent fluid (mud) flowing through the turbine drill to generate the torque that rotates the bit.

Even though the roller hit can operate efficiently at speeds below about 15 orpm, the motor
Diamond bits are used in most turbo drills because they rotate at 0.00 rpm.

Positive motors have a fixed volumetric displacement and speed is directly proportional to flow rate.

Positive motors currently in use or under development are classified into the following three types.

1. MOINEAU motor 2, flexible vane motor 3, sliding vane motor These motors have large capacity displacement and generate large torque at low speed.

The disadvantage of thrust bearings in vertical drilling motors is due to the large dynamic loads caused by bit operation and drill string vibration.

A large oil company placed a recorder at the bottom of the hole and discovered that the dynamic loads often exceeded the bit weight by more than 50 factors.

It was accidentally discovered that when a drill bit weighs 40,000 pounds, the bit bounces off the bottom and creates an excess load of 120,000 pounds.

Such large loads caused thrust bearings to fail in a short period of time, requiring over-design of the bearings.

Two types of bearings were used in vertical drilling motors:

(1) Rubber friction bearings (2) Ball or roller bearings These bearings typically wear out in 20 to 100 hours since they operate in an abrasive drilling mud.

In addition, the comb friction bearing can reduce the output torque by 30% of the turbo drill's output torque.
~40 units will be absorbed by friction.

Extend the life of thrust bearings by balancing bit weight and fluid thrust, taking most of the load off the bearing.
can be increased.

Radial bearings are required on both sides of the drilling motor and on both sides of the thrust bearing.

These radial bearings typically experience lower loads than thrust bearings and therefore have a longer life.

Radial bearings can be divided into the following two types.

(1) Marine Bearings (2) Ball or Roller Bearings Most motors include marine bearings made of brass, rubber, or similar bearings.

Marine bearings are cooled by circulating mud flow.

The rotating seal is the weakest part of the vertical motor.

The improved seal seals the bearing in lubricating oil, increasing its life and allowing it to operate under higher pressure than the bit, greatly increasing drilling rates.

The following six seals have been tested:

(1) Backing seal (2) Face seal (3) Labyrinth seal (4) Radial lip seal (5) Compression seal (friction bearing, marine bearing) (6) Flow rate measurement seal Current drilling motors have a continuous drilling mat. As they pass through the rotating seal, sand and other abrasive particles enter the rotating seal and wear it out over a short period of time.

A turbodrill improvement requires a seal that does not permit even small amounts of leakage, and preferably a means to balance the pressure on the upper and lower seals.

The gist of the present invention is as follows.

A drill motor connected to a string of drill pipes has a rotating shaft that drives a drill bit, such as a rotary bit or a high speed hard diamond bit.

The turbine section is heated between 14 degrees and 2 degrees for maximum turbine efficiency.
It has stationary and rotating blades of crescent-shaped cross section with an exit angle of 3 degrees.

The drill motor may be a positive motor.

The bearing shaft (hairing shaft) is provided with an angle-shaped rotary seal located below a rotary bearing that supports the radial and axial thrust forces.

Lubricating oil is filled from the rotary seal to a predetermined position above the bearing.

A piston seals the lubrication chamber and is pressurized by a perforation mat flowing through the device.

A layer of lubricating oil is above the first piston and has a second piston covering the lubricating oil and transmitting pressure from the drilling mud to the lubricating oil surrounding the bearing.

The drill mud flow branches into a flow for rotating the drill pid and a flow for passing through the drill pid.

The pressure drop across the drill bit is equal to the pressure drop across the drilling motor, which balances the pressure on both sides of the bearing seal.

Embodiments of the present invention will be described in detail below based on the drawings.

First, a turbo drill is shown at 10 in FIGS. 1A to 1D.

The turbo drill 10 is extremely long in the axial direction compared to its diameter, and although a portion is omitted in FIG. 1A, in order to show the whole, FIGS. 1A, 1B, 1C, and Requires Figure 1D.

A typical turbo drill of this type is 7.75 inches in diameter and approximately 20.5 feet long.

Since the turbine section occupies approximately half of the total length of the turbo drill, most of the multi-stage turbine is omitted.

A threaded portion 12 that forms a threaded portion 13 with the lower end of the drill string 14 is provided at the upper end of the turbo drill 10 .

The stationary jacket 11 is provided with a passage 15 communicating with an opening in the drill string 14.

The fixed jacket 11 has a threaded portion 16 that is threadedly engaged with the end of the fixed jacket 19 as indicated by the reference numeral 17 .

The fixed jacket part 18, when assembled,
An annular groove 2 that matches the annular groove 21 provided in the threaded portion 16 of No. 1
has 0.

A locking ring 22 is inserted into the annular space formed by the annular groove 20.21 to prevent accidental separation of the fixed jacket 11.19.

The fixation jacket part 18 is further provided with uniformly spaced holes 23 for applying pressure to the locking song 22 during separation of the fixation jacket 19 from the fixation jacket 11.

Details of this construction are shown in FIGS. 6 and 7 of the US application no. 84.9979, filed Nov. 9, 1977.

The threaded portion is sealed with an O-ring 24 placed in the groove 25.

The turbine section of the turbo drill is arranged in the stationary jacket 19 directly below the joint part 17 that connects with the stationary jacket 11 .

The fixed part of the turbine includes a plurality of stators 26 .

These are filed on November 9, 1977, application number 849.
Figures 3, 3A, 4 and 5 of the '976 U.S. Application
It is explained in detail in the figure.

Stator 26 has an annular structure with vanes or blades 27 .

Stator 26 has an outer surface with slides that match the inner surface of stationary jacket 19 .

Furthermore, the stators 26 are stacked in the axial direction within the stationary jacket 19.

A typical 7.75 inch diameter turbo drill has 50 stators and rotors.

The stator is made of a hard aluminum copper alloy, which is wear resistant and has a slightly higher coefficient of expansion than the rust-resistant stationary jacket 19.

Since the laminated body of the stator 26 expands and is tightly pressed against the inner surface of the fixed jacket 19, it is constantly subjected to a constant compressive force and does not slip.

Further, since it has a high coefficient of thermal expansion, the stator 26 tends to expand more and more due to the high heat generated during use of the turbo drill, and is pressed against the fixed jacket 19, thereby effectively preventing slippage.

The upper end of the stationary jacket 19 is provided with an annular spacer 28 located at the uppermost stator 26 associated with the terminal end of the stationary jacket 11 .

A sleeve 30 is attached to the threaded joint 31 at the lower end of the fixed jacket 19.
The reception is included.

A threaded cylindrical portion 29 is provided on the inner surface.

The lower end of the sleeve 30 is provided with a notch 32 for firmly screwing the sleeve 30 into the threaded joint 31 with a wrench.

When the sleeve 30 is secured in the position shown, its upper end joins the lowermost stator 26 and presses against the stack of stators 26 and the annular spacer 28.

When the sleeve 30 is completely tightened, the laminated body of the stator 26 receives a sufficient compressive force and is brought into pressure contact with the inner surface of the fixed jacket 19, so that no slippage occurs during operation.

The lower end of the cylindrical portion 29 of the fixed jacket 19 is connected to a bearing portion 35.
An annular groove 21a provided in the bearing portion 35 and an annular groove 20 provided in the fixed jacket 19 are directly below the threaded joint 33 that are connected to the upper end threaded portion 34 in the threaded joint 33.
a is located, and the spring locking ring 22a is located there.
should prevent accidental separation.

A hole 23a is provided at the lower end of the fixed jacket for separating the screw joint 33 by applying pressure to the locking ring 22a.

Threaded joint 33 by O-ring 24a provided in groove 25a
Fluid leakage is prevented.

The bearing part 35 extends from the threaded joint 33 to the lower end cylindrical part 36.
The lower end cylindrical part 36 extends to the bearing auxiliary part 3.
It has a threaded joint 37 with 8.

The bearing portion 36 has an inner surface 39 with an inner diameter equal to the inner diameter of the sleeve 30, and a lubrication chamber is formed slightly below the upper end of the bearing portion 35 via a conical surface 41.

Below the conical surface 41 is a conical section 42 that leads to an enlarged inner surface 43 that accommodates various radial and thrust bearings.

The inner surface 43 terminates in a threaded portion on the inner surface of the lower end cylindrical portion 36 of the bearing portion 35 .

A rotating shaft 44 having a cylindrical outer surface 45 located between the upper end of the threaded part 46 and the lower end of the threaded part 47 is provided at the upper end of the turbo drill and inside the fixed jacket 19 .

The rotating shaft 44 has blades or vanes 749 aligned with a plurality of stacked rotors 48 and fixed vanes 27.
has.

Stator 26 includes an outer slip 50 and an inner sleeve 51 having circumferentially evenly spaced blades or bays 727 therebetween.

The outer surface of the outer sleeve 50 is pressed against the inner surface of the stationary jacket 19 to prevent the stator from slipping against the stationary jacket.

The inner surface of the inner sleeve 51 constitutes a bearing surface for smoothly rotating the rotor 48.

The rotor 48 has a bub 52 and a sleeve 53 from which blades or vanes 49 project.

The outer surface 54 of the sleeve 53 forms a bearing surface that mates with the inner bearing surface of the inner sleeve 51 of the stator 26 .

The inner surface 55 of the sleeve 53 and the sleeve 52 has a smooth surface and a groove (for supporting the rotor 48 on the rotating shaft 44).
(not shown).

The blades or vanes are usually crescent-shaped in cross-section.

The vane member 49 has an upper end that is an inlet for receiving fluid, ie mud, and a lower end that is an outlet for discharging fluid.

The shape of the blade or vane is important in the design of a turbo drill.

In particular, the exit angle of the blades or vanes must be set within a very narrow range in order to generate maximum torque within the turbine.

The exit angle of a blade or vane is measured as the angle between the tangent to the curve at the midpoint between the inner and outer curves of vane member 49 and a line perpendicular to the rotor centerline.

The exit angle should be between 14 and 23 degrees, preferably between 18 and 21 degrees.

Maximum rotational torque is obtained at this exit angle.

The structure of vane member 27 of stator 26 is a mirror image of vane member 49.

The rotor 48 is stacked on the rotating shaft 44 together with a vane member 49 arranged together with the vane member 27 of the stator 26 .

The rotor 48 is located on the rotating shaft 44 together with the keyway, and is aligned with the longitudinal groove of the rotating shaft 44 .

A steel wire is inserted into mating grooves (not shown) between the rotating shaft 44 and the rotor 48 to secure the rotor.

The lower end of the rotor stack is connected to the upper end 6 of the spline joint member.
It is in contact with the rotating space ring 64 located at 5.

A cap 66 having a threaded portion 67 on its inner surface and forming a threaded joint with the threaded portion 46 of the rotating shaft 44 is provided at the upper end of the rotating shaft 44 .

When the cap 66 is in the tightened position, its lower end 68 abuts the top rotating member 48 and
The stack of rotating members is strongly pressed onto the rotating shaft 44.

The cap 66 has a central opening 166 and, during operation, the cap 6
6 has one or more threaded holes 69 with set screws 70 to prevent loosening.

At the lower end of the turbine section, adjacent to the lowermost rotating vane, there are one or more openings 235 in the wall of the stationary jacket 19 for discharging fluid.

Optionally, a shield 236 is also provided outside the opening 235.

The spline member 71 has an upper end 65 that abuts the rotating space ring 64 .

The spline member 71 has a threaded portion on its inner surface, and the rotation shaft 44
It forms a threaded joint 72 with the lower end 47 of.

Spline member 71 is hollow and has an outer surface 73 spaced from the inner surface of sleeve 30 to define an annular passageway.

The lower end of the spline member 71 is in contact with an annular spline spacer 76 .

The spline member 71 has a plurality of grooves 77 in the lower end cylindrical portion 75 into which the spline pins 78 are received.

Lower spline member 79 has an upper pin 80 with a groove 81 that receives the other side of spline bin 78 .

Spline 79 has a stepped portion 82 into which the lower end of spacer member 76 is received.

The lower end cylindrical portion 83 of the spline member 79 is connected to the threaded joint portion 8
5 has a threaded portion on the inner surface to receive the upper end of the support shaft 84.

A set screw 86 is provided to prevent the threaded joint 85 from loosening during operation.

The spline member 79 communicates with the opening 166 at one end and with the passage 88 of the support shaft 84 at the other end via the passage 74 of the member 710.

The spline members 71, 79 and the spline bin 78 are
A spline joint between the rotation shaft 44 and the support shaft 84 is formed.

The upper end of the support shaft 84 is connected to the lower end of the spline member 79,
Further, their lower ends are in contact with other bearing sleeves 90, respectively.

The outer surface of the sleeve 89 is spaced apart from the inner surface of the bearing portion 35 to form an annular passage 91, and the annular passage 91 includes:
Lubricating grease or oil and floating piston pair 92.
93 are accommodated.

Piston 92 includes a piston body 94 with a chevron seal on one side and a resilient compressible rail 96 on the other side.

Seals 95 and 96 are compressed by caps 97 held by cap screws.

The seal of piston 92 is of known construction and includes a center spacer and end spacers that compress the seal with cap 97.

Piston 93 has a similar structure to piston 92.

The pistons 92 and 93 are connected to the inner surface of the bearing portion 35 and the sleeve 8.
The lubricating grease or oil is filled between the pistons and below the piston 93.

The bottom of the lubrication chamber 91 includes a sleeve 90 and an annular spacer 98.
It is made up of the upper end of the .

A pair of openings are provided at the lower end of the lubrication chamber 91 to which pipes 99.degree. 100 for supplying lubrication oil to the chamber 91 are connected.

The lower end of the member 98 forms an enlarged step 101 , and the step 101 joins the step 42 of the bearing section 35 .

The lower end of the spacer 98 is joined to the spacer 102 of the bearing section.

A series of radial and thrust bearings are provided below the sleeve 90 and spacer 102 to prevent lubricant leakage at the bottom of the drill by radial seals.

The upper radial bearing consists of an outer ring 103 bearing equally spaced roller bearings 104.

A bearing ring 105 is located on the bearing shaft 84 and completes the radial bearing body.

This type of radial bearing is suitable for the high-speed turbo drill mentioned above.

Radial bearings suitable for 7.75 inch turbo drills are:
McGill Ma nufacturing
Co.

Inc. (Valparaiso, Indiana)
46383) MR-64 and MR-88 related to production
It is a type bearing.

The bearing sleeve 106 is located on the bearing shaft 84 and the bearing ring 1 is mounted to rotate with the bearing shaft 84.
It is joined to the lower end of 05.

Ring 107 is fixed inside bearing portion 35 and is sufficiently spaced from sleeve 107 to allow sleeve 107 to rotate.

The upper end of the ring 10γ is similarly joined to the lower end of a bearing ring 103 fixed within the bearing portion 35.

The lower end of ring 107 is provided with a pair of grooves 108 in which compression springs 109 are housed.

A spring washer 110 is provided for the compression spring 109 and the washer 110 joins the bearing ring 111 of the top thrust bearing.

The thrust bearing includes an upper bearing ring 111, a lower bearing ring 112, and a plurality of evenly spaced bearing rollers 113.
It is composed of.

The upper bearing ring 111 is in close contact with the bearing portion 35 and has a gap with respect to the sleeve 106.

The lower bearing ring 112 is in close contact with the sleeve 106 and has a gap with the inner surface of the bearing portion 35. The spacer ring 114 of one thrust bearing is in close contact with the bearing shaft 84 and has a gap with the bearing portion 35.

The upper end of the spacer 114 has a pair of grooves 11 in which a compression spring 116 that presses the lower bearing ring 112 is located.
5 is provided.

The lower end of the spacer 114 is joined to the sleeve 117 and is provided with a pair of grooves 118 in which a compression spring 119 is housed.

The lower end of spacer 114 also joins the upper ring of the lower thrust bearing.

The lower thrust bearing includes an upper ring 120 that is in close contact with the sleeve 117 and has a small gap with respect to the inner surface of the bearing portion 35 .

There is also a lower bearing ring 121 and equally spaced rollers supported by the bearing race.

The lower bearing ring 121 is in close contact with the bearing portion 35 and has a gap with respect to the sleeve 117.

Immediately below the lower bearing ring 121 is a washer 123 which presses against a compression spring 124 which is supported in a ring 126 and a groove 125 in the upper part.

Ring 126 is identical to ring 107 except in the opposite direction.

An intermediate radial thrust bearing is located below the ring 126 and sleeve 117.

The bearing includes an outer bearing ring 127 that supports bearing rollers 128 for rotational movement.

Inner ring 129 is supported on bearing shaft 84 .

A spacer 130 that is in close contact with the inner surface of the bearing portion 35 is provided below the intermediate radial bearing, and a space ring 131 that has a step 132 that joins with a step 133 of the support shaft is provided. .

The gap between spacers 130 and 131 is sufficient to allow lubricating oil to pass to the lower radial bearing.

Spacers 130, 131 join the top end of the top radial bearing.

This bearing consists of an outer ring 134 and an inner ring 136 with equally spaced rollers 135.

The outer ring 134 is in close contact with the inner surface of the bearing portion 35, and the inner ring 136 is in close contact with the bearing shaft 84.

At the lower end of the bearing portion 35, a bearing auxiliary member 38 is tightened against the lower end of the bearing ring 134 of the lowest radial bearing.

A bearing sealing sleeve 137 is located on the bearing shaft 84 and has one end joined to the lower end of the bearing ring 136 and the other end joined to the ring 138 which is in close contact with the step 139 of the lower end 140 of the bearing shaft. There is.

The bearing auxiliary part 38 has a matching groove 20b including a locking ring 22b.
, 21b, the threaded joint portion is fixed so as not to come off.

The hole 23b is for applying pressure to release the locking ring 22b.

The auxiliary part 38 has a groove 24 in which the O-ring 25b is positioned.
b is a provision.

Dynamic radial seal is auxiliary part 3
8 and the seal sleeve 137 to prevent lubricating oil from leaking from the bearing.

The seal is a chevron seal with upper and lower support rings 141,142.

At the center of the seal is a space member 143.

The upper and lower portions of the spacer member are provided with a plurality of chevron seals that are maintained in a compressed state to prevent lubrication from the bearings during operation of the turbo drill.

The upper spacer member 141 is joined to the locking ring 145,
maintained in that position.

Space ring 142 compresses spring 1 in groove 147
It is connected to 46.

Lower extension 140 is internally threaded as shown at 148 .

This threaded opening receives a hollow connection of a drill bit 150 having a passageway 151.

A turbo drill is shown driving a rotary drill bit 150.

However, as can be readily understood, any type of drill bit can be used.

Particularly effective is a turbo drill equipped with a hard diamond pit as shown in US Pat. No. 3,971,450.

The operation of the drilling device according to the present invention will be explained below.

The turbo drill is assembled as shown in FIGS. 1A-1D.

The device is divided into several parts, which are connected by screw joints.

The jacket of the turbo drill remains stationary and the drill is driven at high speed, applying torque in the direction of disengaging the threaded joint.

In the past, threaded joints were protected by setscrews.

However, the set screws sometimes loosened themselves, making it difficult to obtain the desired protection.

In this construction, the threaded joint is protected by a locking ring.

During assembly of the device, a lubricating oil sufficient to withstand the temperatures generated by the turbo drill is supplied to the lower part of the turbo drill through the lower opening 100.

Lubricating oil penetrates the bearing and radial seal through opening 100.

Similarly, lubricating oil is filled into the space above the piston 93 at the dashed line position by raising the opening 99, thereby raising the piston 92 at the dashed line position.

After the piston has been filled with sufficient lubricant to reach the position shown in solid lines, opening 99.100 is closed.

Once the turbo drill is connected to the drill string 14 (FIG. 1A), the drilling mud is pumped through the drill string 14 (FIG. 1A) at a high flow rate.

After passing through passage 15, the drilling mud flow is divided into flow into passage 166 and flow into the annulus at the upper end of the turbine section.

The drilling mud flows through each turbine section and causes the turbine to rotate at high speed.

One flow of drilling mud flows through each vane 27 of stator 26.
and flows to the vanes 49 of the rotor 48 at high speed.

The shapes of the stator and rotor vanes are as already described in detail.

The shape of the vanes, particularly the exit angle, is designed to provide maximum thrust on the rotor and maximum torque on the rotating shaft.

As already mentioned, the turbine section is formed by a large number of turbine members.

In a typical 7.75 inch turbo drill, 5
0 sets of stators and 50 sets of rotors are used to provide high torque and high speed rotation to the rotating shaft 44.

The drilling mud is discharged from passage 235 through the turbine section and around the turbodrill.

The pressure of the drilling mud at the discharge section is approximately equal to the water pressure inside the hole.

A rotating shaft 44 rotating at high speed is connected to a support shaft 84 via a spline joint.

Some of the drilling mud flows from the turbine section into the annular space around the spline joint at the upper end of the bearing shaft, where it passes through the piston 9.
Apply pressure to the top of 2.

Piston 92 is thus maintained under fluid pressure of the borehole mud flow exiting the turbine section.

The pressure on piston 92 presses on the lubricating oil in space 91, so that, via piston 93, the lubricating oil below piston 93 and around the bearing and radial seal is maintained at substantially fluid pressure.

In the past, floating pistons have been used to pressurize the lubrication system within turbo drills, but drilling mud can erode the piston and penetrate bearing and seal areas, destroying the working parts of the turbo drill. there were.

In the construction of the present invention, a dual piston construction with a fluid lubrication system between the pistons prevents contamination from drilling mud and provides protection and longevity to the seals.

The structure of bearings and seals is extremely important for the operation of turbo drills.

The bearings and seals of turbo drills in the prior art were the most accident prone parts.

With radial bearings, the radial load is considerably smaller than the thrust load, and there are no major space constraints, so there are no major problems, and as mentioned above, roller-type radial bearings are used here. .

Particularly suitable are McGi11's MR-64 and MR-88 type bearings.

The thrust bearing in a turbo drill is an important structural point.

The upper thrust bearing carries the upper thrust during drilling,
The lower thrust bearing carries the load when the motor is stationary.

The preferred thrust bearing is a roller type thrust bearing supported by two annular rings.

The preferred thrust bearing used in the embodiment of the present invention is the Andrews Bearing.
Corp, (Spart anburg jSou
th Carolina 29304) Manufacturing ATH
734 type thrust shaft thrust bearing is 122000
Capable of withstanding dynamic loads of pounds.

As already mentioned, the bearing seal and lubrication system is extremely important.

Bearings in the prior art had extremely short lifespans because they were directly exposed to the drilling mud.

In the improved device, all bearing sections operate with a sealed lubrication system containing lubricating oil pressurized by a floating piston.

Seals that prevent lubricating oil from leaking from the bearings are also extremely important.

Conventionally, backing type seals or compressed rubber seals have been used as seals, but in either case it was necessary to apply pressure to the extent that it was difficult to rotate the rotating shaft.

In this improved device, a multi-stage chevron seal is used as a rotary seal to the bearing to prevent loss of lubricating oil and prevent drilling mud from entering the bearing, increasing the life of the bearing and drill.

The pressure of the drilling mud discharged through passages 166, 74, 88 and 151 into the space adjacent to the drill bit is in equilibrium with the pressure at the bottom of the hole, and that pressure is applied to the bottom seal.

Therefore there is no pressure drop between the seals.

The theory of operation is as follows.

The drilling mud passes through the motor in two streams.

Most of the mud Qt flows through the turbine blades into the annulus above the rotating seal.

The remaining mud Qb passes through the rotor and drill pit.

The rate of flow through the two channels is controlled by the diameter of the nozzle in the pit passage 15151.

The smaller the nozzle, the less mud will flow through the drill.

The fact that the mud is discharged directly from the bottom of the turbine blade into the annulus allows the mud pressure in the annulus to be applied to the floating piston in the lubrication chamber.

Thus, the rotary seal is exposed on both sides to the mud pressure of the annulus, and the mud pressure differential is removed from the rotary seal.

This construction reduces the pump pressure required to operate the turbo drill.

This is because the bit pressure and turbine pressure are connected in series in a normal turbo drill, but in the device of the present invention, they are connected in parallel.

If Q is the total flow rate, then Q=Qt+Qb (1) However,
Qt is the flow rate in the passage 235 of the turbo drill, and Qb is the flow rate in the passage 151 of the drill bit.

The pressure drop across the drill bit and turbo drill is equal since they are both in parallel.

That is, Pb=Pt (2)
Additional pressure resistance may be applied in series with the turbo drill and bit to vary the flow rates of each.

The magnitude of this pressure resistance can be automatically controlled to provide the motor with specific operating characteristics.

For example, it can be controlled as a function of flow ratio, pressure drop, torque, bit weight or rotational speed.

However, 1 is a constant determined by the following equation.

However, W = Mud concentration (lb/ρ4) Qb = Bit flow rate (GPM) C = Nozzle coefficient (70, 95) A = Nozzle area (in2) Turbine pressure drop pt is P t = 2WQ2t (4)
However, K2 = Design coefficient W = Mud concentration (/b/gal) Qt = Turbo drill mud flow tc (GPM) Formula (2)
From (3X4), equation (6) shows that the ratio of flow rates through the turbo drill and the bit is inversely proportional to the nozzle area and is independent of the flow rate.

Therefore, once a bit nozzle is selected, this ratio remains constant.

This simplifies the operation of the pressure balanced turbo drill.

Substituting equation (6) into equation (1), equation (7) shows that when A is O, all mud flows to the turbo drill, and the flow to the turbo drill decreases as A increases. .

A means of protecting the drill hole from erosion is required and this can be accomplished by sloping the outlet passage 235 or by providing a sleeve 236 around the motor to direct fluid upwardly.

Another embodiment is shown in FIGS. 2A-2D.

In this embodiment, a different passage structure is used to balance the pressure on the bearing seals.

The symbols used are substantially the same as those in FIGS. 1A to 1D, and other symbols are used only for different features.

In this embodiment, the opening 235 and the shield 236 in the fixed jacket 19 have been removed.

Spline member 71 has a plurality of passages 63 for allowing fluid from the turbine section of the turbo drill to pass therethrough.

A tube 62 is located within the opening 166 of the cap 66 and is sealed with an O-ring 61.

The tube 62 has a diameter smaller than the diameter of the passageway from which the tube 62 projects, and an annular mud passageway is formed around it.

The bottom of the tube 62 is located inside the upper end of the passage 61 of the drill bit 150.

The tube 62 is sealed in the passage 61 by an O-ring 60.

The tube 62 may be threaded at both ends and secured in place by a threaded joint, in which case no O-ring is used, but a simple seal.

The hollow portion 59 of the lower end 140 of the bearing shaft is connected to the drill bit 15
0 or via the passage 5T of the enlarged portion 140 of the bearing shaft.

If a passageway 57 is used, the passageway 57 is sloped upwardly to prevent hole erosion.

The operation of this embodiment is as follows.

The drilling mud flows through the turbo drill in two streams.

Most of the mud Qt passes through the annular space around the outer periphery of the tube 62 through the turbine blade and the passage 63 of the spline member 71 .

Furthermore, the inner space 59 of the enlarged part 140 of the bearing shaft passes through the bearing part.
and is discharged into the hole through the drill bit passage 58 or the enlarged section 1400 passage 57.

This mud flow flowing through the turbine section causes the turbine to rotate the drill.

Another trickle of mud flows into tube 62 and drill bit 1
50 is discharged from the passage 61 into the hole.

The flow ratio between opening 61 and passage 58 or passage 57 is controlled by the nozzle diameter.

The smaller the nozzle in passage 61, the less mud flow will flow through the drill bit.

The fact that the bottom hem of the hole is discharged from the bottom of the turbine blade shows that the pressure at the bottom of the hole acts on the floating piston of the lubrication chamber.

The bottom lubricating seal is similarly exposed to the pressure at the bottom of the hole.

Therefore, there is no pressure difference between the ends of the rotary seal.

If the total flow rate is Q, then Q=Q t +Q b (1
) However, Qt = flow rate Qb of passage 63 and passage 58 (father is 57) - flow rate of passage 61 The pressure drop in the drill bit and turbo drill part are equal because they are parallel.

Pb=Pt (2) Additionally, a pressure resistor may be connected in series with the turbo drill or bit to change the flow rate ratio.

The magnitude of these pressure resistances can be automatically controlled to provide specific operating characteristics to the motor.

For example, flow rate, pressure drop, torque, bit weight, or
It can be varied as a function of rotational speed.

The pressure drop across the bit is shown as:

However, K1 is a constant expressed by the following equation.

■ Restraint = 1T “Punishment However, W = mud concentration (#/gal) Qb-Bit flow rate (GPM) C=nozzle coefficient (0,95) A = nozzle area (1n2) The pressure drop in the turbine section is expressed by the following equation.

P t = 2WQ2t (4
) However, K2 = Design constant W - Mud concentration (#/gad) Qt = Flow rate of turbo drill part (GPM) From formula (2X3X4) to formula (6), the flow rate ratio is inversely proportional to the nozzle area and is unrelated to the flow rate. It shows that.

Therefore, once a nozzle is selected, this ratio remains constant.

This simplifies the operation of the pressure balanced turbo drill.

Substituting equation (6) into equation (1), the above equation shows that A becomes O
It can be seen that when , the entire flow flows through the turbo drill, and as A increases, the flow in the turbo drill section decreases.

The above structure makes it possible to balance the pressure on the bearing seals of the turbo drill.

The structure of the fluid passage is a positive type (PO3ITIVE DISP
It also works effectively on drilling motors (LACEMENT TYPE).

[Brief explanation of drawings]

FIG. 1A is a partially cutaway front view of the upper part of the turbo drill, showing the structure for applying balanced pressure to both sides of the bearing seal;
FIG. 1B is a partially cutaway front view of a continuous portion of the turbo drill to show the turbine seal, and FIG. 1C is a partially cutaway front view of a further lower continuous portion of the turbo drill to show the seal and bearing. 1D is a partially cutaway front view of the lowest end of the turbo drill to show the joint between the drill motor and the drill bit, and FIG. FIG. 2B is a partially cutaway front view of the upper part of the turbo drill showing the structure of the turbo drill; FIG. Partially cutaway front view to show the continuous part further downward of the turbo drill in Fig. 2B, 2nd
FIG. D is a partially cutaway front view showing the lowermost end of the turbo drill of FIG. 2B. 10 is a turbo drill, 11.19 is a fixed jacket, 35 is a bearing part, 44 is a rotating shaft, 235 is a passage, 92.93 is a floating piston pair, 91 is a lubrication chamber, 150 is a drill bit, 1
41 is a sticker.

Claims (1)

[Scope of Claims] 1. A cylindrical jacket, a rotary shaft supported within the cylindrical jacket and protruding to support a drill bit, a rotary shaft actuated by drilling fluid (drill mud) to support the rotary shaft. a motor disposed within the cylindrical jacket for driving; bearing means within the cylindrical jacket for supporting the rotating shaft; bearing means exposed to the pressure of a drill mat outside the cylindrical jacket; a first rotary seal located between the rotary shaft and the outer sheath below the bearing means, a second rotary seal located between the rotary shaft and the sheath above the bearing means, and a second rotary seal between the two seals. A lubricating oil filled around the bearing means, a first passage communicating from the inlet side of the motor through the motor and the rotating shaft to a point on the outside of the jacket adjacent to the trill hit, and an outlet end of the motor. communicating from an adjacent point through a wall of the jacket to a point outside the wall and exposing the second rotary seal to drill mud pressure outside the jacket;
a second passage associated with a second rotary seal, the first and second passages being connected in parallel;
The pressure drop in the motor section is equal to the pressure drop in the first passage section,
A vertical drilling device configured to connect one end to the lower end of a drill string and the other end to a drill bit, whereby the pressures applied to the first and second rotary seals are balanced. 2. The device according to claim 1, wherein the second passage projects diagonally rearward. 3. The device of claim 1, further comprising an annularly disposed shield at the outlet of the second passage for directing the ejected drill mud to the rear of the jacket. 4. Apparatus according to claim 1, characterized in that it includes a drill bit supported on a rotating shaft and provided with a passageway forming part of the first passageway. 5. A turbo drill including a motor consisting of a stator having a plurality of fixed turbine blades and a rotary i-axis having a plurality of rotating blades that cooperate with the fixed blades to orient the drill mud to drive the rotating shaft. The device according to claim 1, characterized in that: