Patent Application
20050109541
This invention generally relates to rotary jetting tools for drilling and servicing oil and gas wells and production equipment, and more specifically, to a reaction turbine rotor with axially-opposed pressure-balanced mechanical face seals.
There are a wide variety of applications where process or transport tubing becomes fouled with deposits or scale. Water jets, generated by a rotating jetting tool and directed across the internal surface of the tubing or pipe, are commonly used for cleaning these deposits. Such rotating jetting tools can also be used to drill through soil and rock formations. The jet quality provided by the rotating jetting tool is important, especially in harder formations. Jet quality is affected by a number of factors, including standoff distance and upstream flow conditions. Orienting the discharge nozzles of the tool at a large angle relative to its axis of rotation reduces jet standoff distance and improves jetting performance. Uniform upstream flow channels improve jet quality by reducing turbulence intensity. Many designs for rotating jetting tools incorporate relatively small fluid passages, which reduce the pressure and power available for jetting. Other systems require that the operating fluid used be filtered to a high degree, which adds significant expense and complexity. It would be desirable to provide a rotary jetting tool with relatively large flow passages, which does not require the use of an extensively filtered operating fluid.
Rotating jetting tools may use an external motor to provide rotation, or the rotor can be self-rotating. A self-rotating system greatly simplifies the tool operation. In a typical self-rotating system, the jets of liquid are discharged with a tangential component of motion, which provides the torque necessary to turn the rotor. Most self-rotating systems use a sliding seal and support bearing to enable the rotation of the working head. The drawback to this configuration is that the torque produced by the working jets must be sufficient to overcome the static bearing and seal friction. The dynamic friction of bearings and seals is typically lower than the static friction, so once the rotor has started to turn, it can spin at excessive speeds, which can cause overheating or bearing failure. It would be desirable to provide a rotary jetting tool that is configured to prevent such excessive rotation.
Most self-rotating jetting systems also incorporate a thrust bearing to counteract the internal pressure of the fluid against the nozzle. These bearings are subject to high loads and can fail when the rotor's rotational speed is excessive. The thrust load can be eliminated with a balanced or floating rotor design, wherein the shaft is supported by opposed radial clearance seals. If the shaft diameter is the same on both ends of the rotor, there is no thrust due to the internal pressure of the fluid. The clearance seals also act as hydrodynamic journal bearings, which rely upon a thin film of fluid that supports the rotating shaft using hydrodynamic forces. While journal bearings cannot support high thrust or radial loads, they are effective at high velocity—where the hydrodynamic support is greatest.
This approach has been used by Schmidt (as disclosed in U.S. Pat. No. 4,440,242) and Ellis (as disclosed in U.S. Pat. No. 5,685,487) to achieve a self-rotating jet. In the Ellis design, the working fluid is introduced from the tangential surface of the rotor shaft to the center of the rotor by crossing ports. One drawback to this configuration is that the fluid settling chamber is small compared with the sealing diameter of the rotor. In the Schmidt patent, the jet rotor extends well beyond the thrust-balanced section and can be relatively large.
The greatest drawback to the use of radial clearance seals is that clearance seals are prone to jamming with debris, especially when the operating pressure is applied slowly. Sealing, for this approach, is accomplished by maintaining a small clearance, or gap, between the inner and outer elements of the rotor, and leaving a small leakage path for the fluid. Particles approximately the same size or larger than the gap can easily get jammed in the gap and can build up during periods when fluid pressure is low and the rotor is not spinning. When the fluid pressure is increased, such particles are jammed even tighter into the gap and will then prevent the rotor from spinning freely. To avoid this problem, the working fluid must be filtered to remove all particles that might obstruct the smallest gap in the rotor head. Because the gaps must be small to prevent excessive fluid leakage, the fluid must again be filtered to a high degree. In many applications, a relatively large volume of working fluid is required, and filtering the fluid becomes impractical. It is also desirable to be able to pump abrasives or other particles through a jet rotor to enhance the jetting process.
Mechanical face seals overcome the problem of debris jamming the sealing gap. The nominal gap between the sealing surfaces is zero, and leakage is zero when the rotor is not rotating. If fluid is not flowing through the gap, debris cannot be carried into it. Secondly, the sealing gap is not rigidly fixed, as in a radial clearance seal. One element of a mechanical face seal is spring loaded and pressure activated with a secondary seal. If, for some reason, a particle were conveyed into the gap between the sealing faces, the sealing faces can spread, enabling the particle to pass through. Thus, particles are unlikely to become stuck in the sealing gap, and if they do, such particles can escape from the gap as a result of this self-clearing action.
The use of pressure-balanced mechanical face seals for fluid pumping applications is well known in the art. The most common application of mechanical face seals is to provide a fluid seal around a rotating shaft where the shaft penetrates a pressurized vessel so that the fluid is retained in the vessel and does not leak out of the vessel around the shaft. In most cases, such as in single-stage centrifugal pumps, the end of the shaft is exposed to an elevated pressure. This pressure, multiplied by the effective sealing area, produces an end load on the shaft to which a thrust bearing must react. In most pump applications, external support bearings can be provided to withstand the thrust. A mechanical face seal includes a rotating seal ring with a face that slides on a static seal ring. The rotating seal ring is keyed to rotate with the shaft, and is provided with a static seal element that can slide along the shaft. Pressure forces on the rotating element force it axially into contact with a static seal element that is attached to the pressurized vessel. As long as the contact force is greater than the pressure within the pressurized vessel, the seal is effective. The contact force between mechanical sealing faces is determined by the balance ratio of the seal. The balance ratio represents the ratio between the sealed area and the area on which the average pressure between the seal faces acts. This ratio can be adjusted by controlling the seal ring contact area and diameter of the static seal between the rotating seal ring and the shaft. Since the average pressure between the seal faces is normally about one-half the sealed pressure, the seal head will be in equilibrium for a balance ratio of 0.5. It is common practice to choose a balance ratio from 0.65 to 0.75 for contacting face seals. High pressure results in high contact forces between the seal faces, which can lead to premature failure and a high starting torque.
Conventional mechanical face seals have not been used in high-pressure rotating jetting tools for a variety of reasons. The high operating pressure imposes a high shaft end load, which is the product of the operating pressure and the area of the rotating shaft that is sealed. In a conventional design, the shaft load is supported by separate thrust bearings, and the pressure is sealed with a mechanical face seal. The need for separate thrust bearings complicates the tool design and increases the length of the jetting tool. Secondly, the high-operating pressure imposes high contact loads on the seal faces, which results in a high starting torque. The most convenient mechanism for imparting a rotational force to a rotating jetting tool is to use the reaction torque generated by offset jets. This torque is relatively small and is generally insufficient to overcome the friction torque of a conventional mechanical face seal. Finally, it may be desirable to operate rotating jetting tools at relatively high rotational speeds, resulting in a high pressure-velocity (PV) load on any conventional mechanical face seal included within the rotating jetting tool. The PV relationship is defined as the product of contact stress and sliding velocity. High PV values cause premature wear and failure of mechanical face seals. For rotors used in rotating jetting systems for drilling and servicing oil and gas wells and production equipment, an external thrust bearing is impractical, and the thrust loads must be much lower than those induced by the working pressure multiplied by the effective seal area. It would thus be desirable to provide a rotor designed for use in rotating jetting systems for the oil and gas industry that provides the benefits of mechanical face seals, but without the disadvantages of mechanical face seals that were discussed above.
The present invention is a reaction turbine rotor with axially-opposed pressure-balanced mechanical face seals. The rotor is capable of operating with low starting torque, consistent with the relatively low torque generated by the reaction forces of offset jets. The pressure-balanced design of the present invention limits the contact forces on the mechanical face seals, thereby reducing wear and torque. Also, the mechanical face seal surfaces are fabricated from ultra-hard materials, such as tungsten carbide, silicon carbide, and diamond, to minimize wear.
In the event that the rotor contacts the material being cut, the lower mechanical face seal opens and the jetting head is supported by the tool housing, preventing mechanical loading of the seal elements. Contact with the material being cut is accompanied by a predetermined pressure reduction, which can easily be detected on surface, to enable the operator to back the tool off the obstruction. When the tool is backed off, hydraulic features in the tool ensure that the forward face seal will again close and that the tool will restart.
The foregoing aspects and many of the attendant advantages of this invention will become more readily appreciated as the same becomes better understood by reference to the following detailed description, when taken in conjunction with the accompanying drawings, wherein:
Referring to
There are three pairs of dynamic mechanical sealing faces in the rotary jetting assembly of
Preferably, ultra-hard materials are used for each sealing face. Such materials generally having relatively low coefficients of friction and provide superior wear resistance. Polycrystalline diamond surfaces are very resistant to wear, while also providing low frictional resistance to rotation, particularly after an initial period of use (during which the opposed polycrystalline diamond surfaces are subject to mutual smoothing). Other forms of ultra-hard materials may alternatively be employed, such as silicon carbide, cubic boron nitride, and amorphous diamond-like coating (ADLC). Preferably, for each pair of opposed sealing faces, each sealing face is implemented using a different ultra-hard material, which those skilled in the art will recognize provide reduced friction. The opposing faces of a gap between rotor shaft 1 and housing 3 (where the rotor shaft passes through the housing) may incorporate such ultra-hard materials, which act as a radial bushing to maintain alignment between the rotor and the housing.
The present invention reduces startup friction using a unique structure, a mid-face vented mechanical face seal. The mid-face vented mechanical face seal is implemented in seal head 4, which is shown in
Housing 3 includes an orifice 3a disposed immediately distal of lower mechanical face seal 15. Orifice 3a is sized slightly larger than the portion of rotor shaft 1 that passes through orifice 3a, such that a small gap exists between the rotor shaft and the orifice. Because of imperfections in the sealing faces in mechanical face seals, some pressurized fluid will leak past lower mechanical face seal 15 into the gap between rotor shaft 1 and orifice 3a during normal operation. This fluid provides lubrication and a cooling effect on the opposing surfaces of the gap, which act as a radial bushing during normal operation, as noted above. As described in detail below, certain conditions can cause axial movement of rotor shaft 1, resulting in the opening of lower mechanical face seal 15. Under such conditions, more pressurized fluid will flow through orifice 3a than during the normal operating condition. In one preferred embodiment, the gap between orifice 3a and rotor shaft 1 ranges from about 0.003 inches to about 0.0015 inches. The gap provides a leak path for pressurized fluid.
When nozzle head 2 contacts uncut material, or is “set-down,” as illustrated in
Referring to
Pa*A3+Pc*(A2−A3)+Fj+Fc−Pa*(A2−A1)−Po*A1−Fh=0 (1)
where:
The areas and diameters in this analysis are simply a representation of the effective sealing diameters and areas of the seals. These seals have flat parallel faces with constant gap thickness, so the pressure varies linearly from the inner radius to the outer radius. It will be understood that for a given radius, or diameter of the seals, under the condition that a high pressure exists on one side of the radius and low pressure exists on the other, the effective sealing radius, or diameter, is taken to be at the average radius the sealing face.
Assuming Po and Pc are taken relative to Pa, and setting Pa equal to zero, the force balance equation reduces to:
Pc*(A2−A3)+Fj+Fc−Po*A1−Fh=0 (2)
During normal operation the pressure Pc in pressure chamber 12 is equal to the inlet pressure Po. Substituting Po for Pc reduces the force balance equation to:
Po*(A2−A3−A1)+Fj+Fc−Fh=0 (3)
The reaction force for a fluid jet is proportional to the pressure drop across the nozzle (Po) and the nozzle area (Aj). Accordingly, the expression can be rewritten as:
Fj=K*Po*Aj (4)
In one preferred embodiment of the invention, the rotor shaft is held captive between the housing and seal head with equal contact force at the two ends, which implies that forces Fc and Fh are equal. In this case, the equilibrium equation becomes:
A2−A3−A1+K*Aj=0 (6)
The above equation shows that, for a given jetting configuration, if two selected effective sealing areas are chosen, the third sealing area, and therefore the diameter of the third seal, can be calculated to produce any desired contact force between the stationary and rotating elements. In a preferred embodiment, diameter D3 is maximized to reduce the flow velocity, pressure differential, and turbulence into reservoir 20 of nozzle head 2. Diameter D2 is made larger than diameter D3, within geometric constraints of the system. Diameter D1 is then sized to produce a light contact load on the lower seal when the largest expected nozzle combination is used.
Referring to
Fh+Pa*(A2−A1)+Pc*(A5−A2)−Fs−Po*(A4−A1)−Pa*(A5−A4)=0 (7)
Making similar assumptions as before, the force balance equation reduces to:
Fh−Fs+Po*[(A5−A2)−(A4−A1)]=0 (8)
The contact force between the seal head and rotor shaft is then:
Fh=Fs+Po*[(A4−A1)−(A5−A2)] (9)
The values of A1 and A2, and therefore, D1 and D2, are determined as described above to balance the forces on the rotor shaft. The values of A4 and A5, and therefore, D4 and D5, can be selected so that the contact force is proportional to the working pressure, and the constant of proportionality can be positive, zero, or negative. These diameters are selected to impart a small positive force, Fh, as a function of pressure, so that seal head 4 and rotor shaft 1 remain in contact. By careful selection of these diameters, the contact force can be kept small enough that the torque produced by the fluid jet(s) can overcome the static friction torque from the contact between rotor shaft 1 and housing 3, as well as from the contact between rotor shaft 1 and seal head 4.
If rotor shaft 1 were allowed to spin unrestrained at full pressure, the rotation speed would be very high, causing excessive wear of the sealing components. To prevent this problem, a braking apparatus is included in one preferred embodiment of the present invention, as explained below. Referring to
In one preferred embodiment of the invention, the rotary jet head is protected by a circular gage ring 30 that is coupled to housing 3. The gage ring is forced into contact with the formation to be drilled or material to be removed from a tube. Coiled tubing and jointed tubing systems are commonly lowered or pushed into a well with a system that is equipped to monitor the force on the working end of the tubing. When the force rises, the operator knows that the tool is in contact with the formation in the borehole. The gage ring prevents any further advance of the tool until all of the material ahead of the gage ring is removed. This approach enables drilling of a near gage circular hole in rock. Gage ring 30 also generally protects nozzle head 2 from coming into contact with the formation. In the event that the applied force is too high, the rotating head may contact the formation anyway. When nozzle head 2 contacts the formation, it will be pushed back, and the back face of nozzle head 2 will come into contact with housing 3 (i.e., gap 22 will be eliminated by the movement of nozzle head 2). The axial movement of nozzle head 2 and rotor shaft 1 causes lower mechanical face seal 15 to leak. This leakage is accompanied by a loss of fluid pressure when pumping fluid at a fixed flow rate. The operator thus has an indication that the rotor head has contacted the formation and stalled. The force on the tool may then be reduced or the tool may be pulled away from bottom of the borehole to address the problem.
The embodiment described above achieves the vented upper mechanical face seal by forming an annual recess in the seal head. An alternative embodiment achieves a similar vented upper mechanical face seal by forming an annular recess in the proximal face of the rotor shaft. This latter embodiment is schematically illustrated in
In another embodiment of the present invention, a rotary jetting tool includes a pressure-balanced lower mechanical face seal configured to reduce a startup torque required to initiate rotation of the rotor and nozzles. The embodiments described above have reduced the startup torque required by using a vented upper mechanical face seal, which results in an area of low pressure being disposed proximate a proximal end of the rotor. This lower pressure area above the rotor reduces a startup torque required by reducing the force exerted by the operating fluid on the rotor. A similar reduction in the startup torque can be achieved by pressure balancing the lower mechanical face seal, instead of by venting the upper mechanical face seal. Pressure balancing the lower mechanical face seal to reduce startup torque is a accomplished by providing a volume of relatively high pressure in fluid communication with the lower mechanical face seal. This volume of relatively high pressure will in part counteract the force exerted on the rotor by the column of working fluid disposed proximal of the rotor. In short, the column of working fluid above the rotor provides a force that loads the lower mechanical face seal. This force can be offset in part by providing a volume of relatively lower pressure adjacent to the upper mechanical face seal, or by providing a volume of relatively high pressure adjacent to the lower mechanical face seal.
An upper mechanical face seal 16b is achieved between a distal face of seal head 4b and a proximal face of rotor shaft 1b. A lower mechanical face seal is achieved between a distal annular face of rotor shaft 1b and housing 3. Annular recess 13b separates the lower mechanical face seal into an inner lower mechanical face seal 15a and an outer lower mechanical face seal 15b. As discussed above, ultra-hard surfaces can be used to implement each sealing face, and it is particularly preferred that each face in a sealing face pair be implemented using a different type of ultra-hard art material.
In the embodiment illustrated in
Although the present invention has been described in connection with the preferred form of practicing it and modifications thereto, those of ordinary skill in the art will understand that many other modifications can be made to the present invention within the scope of the claims that follow. Accordingly, it is not intended that the scope of the invention in any way be limited by the above description, but instead be determined entirely by reference to the claims that follow.
This application is based on a prior copending provisional application Ser. No. 60/520,919, filed on Nov. 17, 2003, the benefit of the filing date of which is hereby claimed under 35 U.S.C. § 119(e).
| Number | Date | Country | |
|---|---|---|---|
| 60520919 | Nov 2003 | US |
