mud pump discharge module free sample

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The invention relates to an assembly for quickly securing and releasing a component to a pump housing and more particularly to a retainer assembly for releasably mounting a piston liner within a hydraulic cylinder on the module of a pump.

Heavy duty large horsepower pumps are used to pump fluids or slurries with entrained solids. In the oil industry, for example, slush or mud pumps are used to pump viscous fluids, such as drilling muds, cement, or other well fluids. Although mud pumps may be either centrifugal or reciprocating type pumps, typically mud pumps are reciprocating pumps using one or more pistons and hydraulic cylinders with liners to generate the high pressures required to pump these viscous fluids in and out of the well.

Mud pumps include a fluid end and a power end. In the fluid end of one type of a triplex mud pump, for example, there are three sets of suction modules and discharge modules in fluid communication. A suction manifold is connected to the fluid inlets of the suction modules for receiving fluids and passing those fluids to each of the suction modules. A discharge manifold is connected to the fluid outlets of the discharge modules for discharging the pumped fluids. Each module encloses a set of flow passages with check valves for controlling the direction of flow of the fluids. A check valve is disposed at the suction module fluid inlet to only allow fluids to enter the suction module inlet end of the module and another check valve is disposed at the discharge module fluid outlet to only allow fluids to exit the the discharge module for flow into the discharge manifold.

Each discharge module includes a liner retainer flange attached to the discharge module. The liner retainer flange attaches to a replaceable liner within which a pump piston reciprocates. The piston is a generally cylindrical steel member having a polymer, such as polyurethane, bonded to its outer diameter for sealingly engaging the inner cylindrical wall of the liner to ensure a fluid tight seal required for drawing the low pressure fluids through the suction manifold and module flow passages. The seal integrity must be maintained to withstand the high discharge pressure on the discharge stroke. The power end contains the gears that reciprocate the pump piston within the liner for pumping the fluid through the module passages in the fluid end and thence out the discharge valve.

In operation, on the suction stroke, the pump piston draws fluids through the suction manifold and suction valve as the piston strokes within the liner. On the discharge stroke, the check valve in the discharge module opens simultaneously as the suction valve closes preventing suction back flow into the suction module. Fluid in the liner is compressed and pressure is built up until the pressure overcomes well bore pressure so as to pump the mud into the well. The piston then reverses for another suction stroke whereby the check valve in the suction module opens and the discharge valve closes simultaneously, the piston now making a suction stroke.

As the piston reciprocates within the liner, friction wears the liner. Further, the fluid passing through the fluid end includes particulates and other solids which wear away and destroy the liner and piston. When the liner and piston degrade, the fluid seal is lost and the pump becomes much less efficient. Also, the reciprocation of the piston in the liner causes pulsations that over time cause the liner to become loose within the containment of the liner retainer flange thus resulting in a degradation of the seal at the face of the liner and the seal at the face of the liner wear plate. Therefore, it is important to be able to replace the liner as a part of routine maintenance (or when emergencies occur from seal failure while drilling) to ensure that the pump operates efficiently and can control well pressure. It is also important to have a means for fastening the liner to the liner retainer flange so as to ensure that the liner remains firmly secured despite extended reciprocation of the piston assembly within the liner.

Typically each liner retainer flange, and the cradle of the pump power end are all secured to the fluid end module by studs and threaded connections. Because of the environment in which the mud pump operates and the corrosive nature of the fluids being pumped, the studs and threaded connections, such as nuts, become corroded and are difficult to unthread for the replacement of the liner. Often, the threaded connections have been over tightened, making it even more difficult to unthread. Where the liner is retained by an end cap, a steel bar is inserted into a guide hole in the side of the end cap and then the cap is unscrewed using a significant amount of torque. This end cap is very heavy as it must have sufficient strength to keep the liner from moving, even with pressures up to 7500 psi. Where a nut or end cap resists unscrewing, a sledge hammer is used to hammer on a socket wrench or a special hammer wrench is used to loosen the nut or cap. Such activity is obviously dangerous. In some regions of the world local laws prohibit the use of sledge hammers for personnel safety reasons or to avoid the risk of an explosion due to sparks.

Prior art liner retention systems include spring mechanisms around each stud with an end flange for securing the liner against a fluid end module. Hydraulic pressure is applied to the spring mechanism of each stud by a small hydraulic pump to remove the clamping force of the spring mechanism. The release of the clamping force allows the removal of the clamping flange of the liner retention system. Individually actuated spring loaded studs cause an uneven pressure to be applied to the clamping flange. Further, the clamping force is limited because of the limited space available to hold numerous springs.

The liner retainer assembly of the present invention includes a liner retainer flange that is mounted on the discharge module of the fluid end of a pump. A pressure actuated hydraulic clamping piston with related actuated, conical dished washers and necessary static and sliding seals is disposed within the retainer flange . The hydraulic pressure actuated clamping piston is configured to receive and hold the liner. The hydraulic clamping piston and an end cap maintain the liner in contact with the module during actuation. The hydraulic clamping cylinder includes a counterbore which is divided by the hydraulic piston into a fluid cavity and a spring cavity. The spring cavity houses a plurality of springs which bias the hydraulic piston, end cap, and liner towards the module, thus providing a strong clamping securing force when the hydraulic pressure is released. The fluid cavity communicates with a supply of hydraulic fluid for biasing the hydraulic piston away from the module to activate the springs. By pressurizing the fluid cavity, the springs are compressed so as to disengage the liner retaining end cap from the liner and allow the unthreading of the liner end cap to then remove the liner.

The liner retainer assembly permits preloading or prestressing of the liner against the module of the fluid end of the pump so that the liner will not loosen upon the reciprocation of the pump piston within the liner. Further, the liner may be easily secured and unsecured from the module without the necessity of a sledge hammer or other methods for applying excessive amounts of torque to a securing fitting. The assembly of the present invention permits the easy and quick replacement of the liner as necessary.

Referring first to FIGS. 1 and 2, there is shown a fluid end module 10 and a cradle 28 of the pump power end. The pump is of the type used to pump fluids, such as drilling muds, cement or the like. Pumps of this type are well known. A wear plate 14 defines a bore 16 which leads into liner 20. The module 10 is used for the transfer of fluid from the suction manifold and suction module (not shown) to the discharge manifold (not shown) and discharge module.

An exemplary liner retaining flange assembly 18 of the present invention is used to secure liner 20 within a hydraulic cylinder 30 mounted on module 10 and liner retainer flange 22. Those of skill in the art will understand that a pump piston (not shown) attached to the power end of the pump is reciprocated within the liner 20 to effect the desired pumping action to flow fluid through the fluid end module 10 of the pump. Hydraulic cylinder 30 provides an open end into which the liner 20 is inserted. Module 10 also provides a counterbore 12 for the adjacent wear plate 14 against which it is desired to retain the liner 20 during operation of the pump piston. It can be appreciated that the purpose of wear plate 14 is to avoid the end of liner 20 wearing module 10 due to the reciprocation of the piston within liner 20. However, wear plate 14 may cause wear to the module 10 if the liner 20 is not securely affixed. Wear plate 14 may be replaced should that wear become excessive. It is noted that the end of the liner 20 adjacent the wear plate 14 includes an internal annular groove 59 with seal member 61 for sealingly engaging the wear plate 14 and the other open end 15 of liner 20 includes an external annular load-bearing shoulder 60 which retains end cap 64 (FIG. 2).

The hydraulic cylinder 30 includes a threaded, reduced diameter portion 24 and an enlarged diameter portion 32. Reduced diameter portion 24 is secured in a threaded or splined relation at 23 to liner retainer flange 22 that is located in an abutting relation to the module 10. Bolted studs 26 secure the cradle 28 of the pump power end, the liner retaining flange 22 and hydraulic cylinder 30 to module 10.

When these components are assembled, the annular flange 44 of piston 42 forms hydraulic cavity 58 and outer spring cavity 35. The hydraulic fluid port and fitting 38 communicates with hydraulic cavity 58 for applying hydraulic pressure to flange 44. The spring cavity 35 houses a plurality of axially compressible Belleville springs or washers 56. Retainer ring 48 has external threads 50 which threadingly mate in a complimentary fashion with the internal threads 34 of enlarged diameter portion 32. The washers 56 bear against the retainer ring 48 and annular flange 44. Enough springs are used so as to insure sufficient force is generated to prevent movement of liner 20 when pump pressure is at maximum. The retainer ring 48 secures the washers 56 and hydraulic piston 42 within the enlarged diameter portion 32 of hydraulic cylinder 30. O-ring 49 provides a fluid-tight seal between the piston 42 and enlarged diameter portion 32.

Referring particularly to FIG. 2, the retaining assembly 18 is shown ready to secure the liner 20 in place. A hydraulic hose 62 is secured to the external port and fitting 38 for supplying hydraulic fluid to inner hydraulic cavity 58. As fluid pressure is supplied to cavity 58, fluid pressure is exerted against flange 44 urging piston 42 outward toward retainer ring 48. As annular flange 44 of piston 42 is so moved, springs 56 are axially compressed. As springs 56 are compressed, the threaded end 46 of piston 42 extends further away from wear plate 14 and module 10.

Referring now to FIG. 3, the retaining assembly 18 is shown completely assembled with the liner 20 securely affixed within the hydraulic cylinder 30. Once end cap 64 has been affixed, the fluid within the hydraulic cavity 58 is evacuated through port and fitting 38 permitting the springs 56 to bias flange 44 toward wear plate 14 and module 10 and bias end cap 64 against the other end 15 of liner 20. As the hydraulic pressure in the hydraulic cavity 58 is released, stored energy from the compression of springs 56 is released to load the liner 20 longitudinally. As a result, the energy stored by compressing springs 56 is transmitted to the liner 20 in order to load it longitudinally against wear plate 14 and module 10.

It is noted that the arrangement of the present invention permits a liner to be replaced rapidly and easily and without the use of extra tools or having to apply excessive torque. Further, a prestress force is applied to the liner 20 so that it is longitudinally compressed against wear plate 14 and module 10. This load or prestress securely holds the liner 20 against the wear plate 14 despite repeated reciprocation of the pump piston within liner 20.

In operation as shown in FIG. 4, the liner 20 is installed and removed in a manner similar to that described with respect to liner retaining assembly 18. Fluid is introduced into each individual fluid chamber 82 urging the associated piston 74 to move toward the open end 15 of liner 20. Energy is stored through axial compression of the Belleville springs 94. The end cap 98 is placed onto the liner open end 15 so that the shoulder 104 engages shoulder 60 of liner 20. Nuts 86 are then tightened onto each piston 74. Again, the nuts need only be hand tightened. Fluid is then evacuated from the fluid chambers 82 and the Belleville springs 94 bias pistons 74 toward the wear plate 14 and the module 10, thus loading liner 20 longitudinally against wear plate 14 and module 10.

This application is a continuation of pending U.S. patent Ser. No. 12/220,876 entitled “A Reinforced Smart Mud Pump” which take priority to provisional application for patent filed Jul. 30, 2007 bearing Ser. No. 60/962,637 and is incorporated by reference herein as if fully set forth.

This invention relates generally to the field of mud pumps and more specifically to a reinforced smart mud pump. Mud pumps that use piston displacement, produce imposed forces that cause wear and tear on various pump components, including pump cross head piping, cylinders, inlet and discharge valves, seal components including piston or plunger seals, the pump cylinder block or so-called fluid end, and other components. There has been a need to provide increased longevity and performance for such pumps and to determine if deteriorations in pump performance are occurring, to analyze the source of decreased performance and to further real time control and data to monitor and in some cases change the operating characteristics before damage occurs to the pump. The use of greatly strengthened components in combination with a computer controlled system integrated with a real time monitored and controlled reset relief valve may be integrated into an oilfield application to prevent catastrophic pump failure and extend pump life.

Pump operating characteristics often have a deleterious effect on pump performance. For example, delayed valve closing and sealing can result in loss of volumetric efficiency. Factors affecting pump valve performance include fluid properties, valve spring design and fatigue life, valve design and the design of the cylinder or fluid end housing. Delayed valve response also causes a higher pump chamber pressure than normal which in turn may cause overloads on pump mechanical components, including the pump crankshaft or eccentric and its bearings, speed reduction gearing, the pump drive shaft and the pump prime mover. Moreover, increased fluid acceleration induced pressure “spikes” in the pump suction and discharge flowstreams can be deleterious. Fluid properties are also subject to analysis to determine compressibility, the existence of entrained gases in the pump fluid stream, susceptibility to cavitation and the affect of pump cylinder or fluid end design on fluid properties and vice versa.

Still further, piston or plunger seal or packing leaking can result in increased delay of pump discharge valve opening with increased hydraulic flow and acceleration induced hydraulic forces imposed on the pump and its discharge piping. Moreover, proper sizing and setup of pulsation control equipment is important to the efficiency and long life of a pump system. Pulsation control equipment location and type can also affect pump performance as well as the piping system connected to the pump

In prior art, the control of a mud pump has been disclosed focused on piston position for acquiring information about the pump and its performance characteristics. For example, U.S. Pat. No. 6,882,960 to Miller, shows a system for monitoring and analyzing performance parameters of reciprocating piston, or power pumps and associated piping systems. This patent fails to disclose the innovative aspects of the present invention.

Nothing in the prior art shows a computer integrated mud pump with significant strengthening features that increase the life cycle of a pump in the manner of the present invention with real time control of significant operating functions and feedback from various sensors and reset relief valves.

Another advantage of the invention is to provide a mud pump that utilizes transducers in line with the ambient pressure in conjunction with a computer controlled pressure relief valve to record and monitor pump characteristics and control the pump to prevent catastrophic failure.

Another advantage of the invention is to provide a mud pump that transmits data to a computer for later analysis of important operating characteristics.

A further advantage of the invention is to provide a mud pump that can be controlled during its operation to prevent certain damaging events to the pump or underlying pressurized system.

In accordance with a preferred embodiment of the invention, there is shown a pump system for movement of fluids having a reciprocating piston power pump having at least three reciprocating pistons operable to displace fluid from a housing having a pumping chamber, an integrally forged crankshaft operably connected to the pistons, at least one sensor operable to sense ambient conditions on the pump, and a computer control for processing data from the sensor to regulate the operation of the pump in response to the data.

In accordance with a preferred embodiment of the invention, there is shown a pump system for movement of fluids having a reciprocating piston power pump having at least three reciprocating pistons operable to displace fluid from a housing having a pumping chamber, an integrally forged crankshaft operably connected to the pistons, polyflorocarbon infused treatement applied to at least one crosshead and one crosshead slide in the system, and at least one sensor operable to sense ambient conditions on the pump.

In accordance with a preferred embodiment of the invention, there is shown a pump system for movement of fluids having a reciprocating piston power pump having at least three reciprocating pistons operable to displace fluid from a housing having a pumping chamber, an integrally forged crankshaft operably connected to the pistons, at least one sensor operable to sense ambient conditions on the pump, a computer control for processing data from the sensor to regulate the operation of the pump in response to the data; and pressure sensors for monitoring fluid pressure operably connected to the computer control. In addition, upper and lower limits to temperature, vibration and pressure can be set. Further, in a preferred embodiment, all lubrication and water pumps must be on before the control system will permit unit to be operated.

In a preferred embodiment of the present invention, there is shown a reciprocating plunger or piston power pump. The pump includes additional features not found in conventional reciprocal pumps as heretofore described. The basic operation of the pump is similar to a triplex plunger pump configured to reciprocate three spaced apart plungers or pistons, which are connected by suitable connecting rod and crosshead mechanisms, to a rotatable crankshaft or eccentric. FIG. 1 shows a side view of the pump and auxiliary equipment according to a preferred embodiment of the invention. Pump housing 100 covers the internal pump components and allows for a variety of conventional means to move the rotatable crankshaft 114 such as an electric motor 102 and belt drive 104. Within pump housing 100 may be disposed a pressurized lubrication spraying system that continuously feeds lubricant such as oil around the crankshaft and associated internal pump components.

A suction module 104 of conventional design houses each plunger which operates each section of the triplex plunger pump. A pressure relief valve mount 110 allows for attachment of a reset relief valve (show in FIG. 5) to at least one suction module 106 in a system for pumping drilling mud composed of water, clay and chemical additives, down through the inside of a drill pipe of an oil well drilling operation. The drilling mud is pumped at very high pressure so the mud is forced out through a bit at the lower end of the drill pipe and returned to the surface, carrying rock cuttings from the well. In this illustrative example the drilling mud from the pump system is fed into the attached vessel 112 sometimes known as a dumping ball. In a preferred embodiment all components are installed on a full unitized skid 120 providing room for motors, starters, sheaves, belts and any associated test equipment and a solid platform for installation of bracing 122 for component parts. Vessel 112 smoothes out the pulsation caused by the pumping action of the system to deliver pumped fluid out of port 115 in a more controlled manner.

With the use of computer modeling, high technology engineering, metallurgical and mechanical enhancements, the smart pump is revolutionary in design. The pump is manufactured using advanced materials and techniques including an integrally forged and balanced crank. This provides significant strength advantages and increases the life cycle of the pump. Unlike prior art pump systems, the crank is not a porous unbalanced crank casting, nor is it a fabricated with separate plates and bars and later welded together. As shown in FIG. 2 the crank 200 is fabricated from an integrally forged single ingot by the open hammer forging process resulting in a single piece with no welding or pinning of multiple pieces. A connecting rod 202 for each plunger is attached to the crank 200 such that the crank freely rotates and creates a reciprocating action in each connecting rod 202 attached to the crank 200. A crosshead 204 and crosshead slide 206 attach to each connector rod to help retain engine oil in the crankcase. A plunger connection 210 at each connector rod 202 end provides a means of attachment for a variety of rod and plunger components.

All of the ground crossheads 204 and honed crosshead slides 206 are treated with a polyflorocarbon coating for more lubricity and wear performance characteristics resulting in 15% less energy cost. This process has been used in the racing industry with much higher performance results. The poly-fluorocarbon coating is applied to each crosshead and slide and helps retain engine oil on the component surfaces during intense heat and extreme pressure. The oil is essentially absorbed into each crosshead 204 and slide 206 in such a way as to increase their lubricity. The crosshead 204 center line alignment is laser monitored for even wear. In a preferred embodiment of the invention the pump uses suction modules, discharge modules and discharge manifold and a double helical gear set. All gear sets are prepared with mesh test providing contact tapes accompanied by digitals to AGMS 11+ standards. All gear sets are further preferably chemically treated to 0.4 RMS or better to exponentially increase bearing life.

In the present invention, the crankshaft is a integrally forged piece to increase its operating strength significantly, and is an integral part of the crank 200. The terms crank 200 and crankshaft are used interchangably although in conventional pumps and engines the crank 200 and crankshaft may comprise separate components. Turning again to FIG. 1 we see the crankshaft includes a rotatable input shaft portion adapted to be operably connected to a suitable prime mover, such as an internal combustion engine or electric motor 102, as an exemplary installation. The crankshaft is mounted in a suitable, so-called power end housing 114 which is connected to a fluid end structure configured to have a plurality of pumping chambers, in this example, three separate pumping chambers exposed to their respective plungers or pistons. Plungers refer to the rod, rod joints and piston end portions of the plunger unless otherwise specified.

FIG. 3 shows a partial cutaway side view of the pump according to a preferred embodiment of the invention. The fluid end comprises a housing having the series of plural cavities or chambers for receiving fluid from an inlet manifold by way of conventional poppet type inlet or suction valves contained each suction module 300. The piston or plunger 304 projects at one end into the chamber and is connected to a suitable crosshead mechanism 306, including a crosshead extension. The crosshead extension is operably connected to the crankshaft using connecting rods 308 as described above. Each plunger 306 projects through a conventional packing 310 or plunger 306 seal. Each chamber for each plunger 306 is operably connected to a discharge piping manifold by way of a suitable discharge valve. Valves may be of a variety of conventional designs and are typically spring biased to their closed positions. Valves may also include or are associated with removable valve seat members. Each valve may also have a seal member formed thereon engageable with the associated valve seat to provide fluid sealing when the valves are in their respective closed and seat engaging positions. A unique feature of a preferred embodiment of the the smart pump is a positive air pressure plunger seal 312 to prevent leakage and prevent wear on the plunger. FIG. 3A shows a cross sectional view of the positive air pressure plunger seal about a synchronized plunger according to a preferred embodiment of the invention. As the plunger 350 moves reciprocally, pressurized air is introduced into through seal 352 through a plurality of pressurized air ports 354 preventing fluid leakage from the plunger 350 end which engages and moves drilling mud through the aforementioned valves. The pressurized air flow 356 isolates any material from scoring and etching the plunger 350 and reduces friction between the seal wall and the plunger 350.

In a preferred embodiment the pump may be fitted with a P-QUIP ® fluid end systems including P-QUIP ® kwik clamp liner retentions system, P-QUIP ® kwik rod system and cover system. The P-QUIP ® kwik-clamp valve and strainer cover retention system has a very fast and safe access with much reduced down time and LTAs (Lost Time Accidents) due to mishaps. Hammers or cheater bars are not required and the system includes an automatic clamping means which results in no more under or over tightening—caps will not loosen off in use. This all results in easy installation and easy operation with conventional air or hand operated hydraulic pump. Like the Liner Retention System, the valve and strainer covers are sealed firmly in the fluid end by means of a spring mechanism. Similarly, the outer cover is removed when the clamping force is released by means of the hydraulic pump. This system heightens safety as it is no longer necessary to tighten OEM type threaded cap retainers by hammer.

The P-QUIP ® kwik rod system allows for fast and safe piston changes and is constructed of 17.4 PH martensitic high-resistance stainless steel, eliminating corrosion. Three piece rod system is held together using pins, which eliminates prior systems and clamp type systems. The P-QUIP ® kwik cover system allows for fast and save valve access changes with no hammers or cheater bars. The cover offers a positive retention force against the plug eliminating the need for retightening. The P-Quip ® kwik rod pump rod system permits fast and safe piston changes resulting in reduced down time. There is no need for heavy clamps or connecting threads and studs. The clamping force is automatically controlled resulting in no broken rod ends. Due to its construction, its self alignment facility gives improved piston and liner life. There are no corrosion problems as all parts are stainless steel, including hard-surfaced stainless steel power end rods. It has an integral liner flushing systems.

On conventional mud-pump rod systems, the rod is held together by means of taper clamps and screw threads which are slow and awkward to assemble correctly and readily wear out. Due to their design, the clamps obscure vision of the rod joints, preventing a check being made that the rod is correctly aligned. Uneven loads are imposed on the flanged rod ends, resulting in premature failure.

The P-QUIP ® Kwik-Rod system avoids these problems as the rod components are held together by powerful spring-loaded ends on a release link in the center of the rod assembly. The ends of the release link are attached to the pony rod and piston rod by means of high tensile stainless steel pins held in shear. The shear force is very quickly and easily released by a few strokes of a small hydraulic pump.

Dismantling and re-assembly of the complete rod system takes under one minute. Furthermore, there are no flanged joints on the rods to chip, wear or break. Hence, rod life is enhanced and, because rod alignment is readily achieved, significantly improved swab and liner life is generally obtained. FIG. 4 shows a partial cutaway overhead of the pump housing according to a preferred embodiment of the invention. Plunger rods 400 are synchronized based on the crank speed and can be checked for alignment as described herein. In a preferred embodiment, a viewing port 402 or access cover allows access to plunger rods 400 for alignment and maintenance.

Further enhancement to the pump is achieved by use of ductile iron for the crossheads, the crosshead slides and the connecting rods. The connecting rods are solid ductile iron which reduces their elongation during setoff to zero. The typical rod experiences stretching or elongation during set-off when the relief valve is activated. The benefit of ductile iron in increased strength and higher tensile strength. All the major components, the crank, the crossheads, the slides and the rods by use of mettalurgical engineering are designed to bring the pump tensile strength up to 200,000 p.s.i.

The pump is network and web based with a data acquisition system to monitor pump performance and constantly evaluate pump valve dynamics. Pressure transducers are located in pump chambers used to determine valve sealing delays, fluid compression delays, chamber overshoot pressure, crosshead loading shock forces and chamber volumetric efficiency. Pressure transducers are also located in suction piping and manifolds and discharge piping and manifold. Temperature is similarly monitored for fluid temperature for mud properties and power end lubrication. Further, there is real time power input data to calculate system mechanical efficiency.

Those skilled in the art will recognize that the present invention may be utilized with a wide variety of single and multi-cylinder reciprocating piston power pumps as well as possibly other types of positive displacement pumps. However, the system and method of the invention are particularly useful for analysis of reciprocating piston or plunger type pumps. Moreover, the number of cylinders of such pumps may vary substantially between a single cylinder and essentially any number of cylinders or separate pumping chambers and the illustration of a so called triplex or three cylinder pump is exemplary.

The performance analysis system of the invention is characterized, in part, by a digital signal processor which is operably connected to a plurality of sensors via suitable conductor means well known in the art. The processor may be of a type commercially available as previously described and may wireless remote and other control options associated therewith. The processor is operable to receive signals from a power input sensor which may comprise a torque meter or other type of power input sensor. Power end crankcase oil temperature may be measured by a sensor. Crankshaft and piston position may be measured by a non-intrusive sensor including a beam interrupter mountable on a pump crosshead extension for interrupting a light beam provided by a suitable light source or optical switch.

A vibration sensor may be mounted on the power end or on the discharge piping or manifold for sensing vibrations generated by the pump. Suitable pressure sensors are adapted to sense pressures at numerous locations, including the inlet piping and manifolds. Other pressure sensors may sense pressures in the pumping chambers of the respective plungers or pistons. Other pressure sensors sense pressures upstream and downstream of a discharge pulsation dampener. Still further, a fluid temperature sensor may be mounted on discharge manifold or piping to sense the discharge temperature of the working fluid. Fluid temperature may also be sensed at the inlet or suction manifold.

Pump performance analysis using the system may require all or part of the sensors described above, as those skilled in the art will understand and appreciate from the description which follows. The processor may be connected to a terminal or further processor, including a display unit or monitor mounted in a housing connected to the pump system and main housing. Still further, the processor may be connected to a signal transmitting network, such as the Internet, or a local network.

FIG. 4A shows a schematic drawing of a representative computer controlled monitoring system. A computer 450 receives input from from sensors 452 and modules 454 in the pump system and uses software algorithyms to analyze pump performance, record and display operational data on a visual screen display 456. The computer controlled monitoring system may be adapted to provide a wide array of graphic displays and data associated with the performance of a power pump on a real time or replay basis. A substantial amount of information is available including pump identification (Pump ID) crankshaft speed, fluid flow rate, time lapse since the beginning of the display, starting date and starting time and scan rate. The display 456 displays discharge piping operating pressure, peak-to-peak pressures, fluid flow rate induced peak-to-peak pressure, fluid flow induced peak-to-peak pressure as a percentage of average operating pressure, pump volumetric efficiency and pump mechanical efficiency. The display 456 also indicates discharge valve seal delay in degrees of rotation of the crankshaft from a so called piston zero or top dead center (maximum displacement) starting point with respect to the respective cylinder chambers of the pump, as well as piston seal pressure variation during fluid compression and suction valve seal delay in degrees of rotation of the crankshaft or eccentric from the top dead center position of the respective cylinder chambers. Still further, the pump type may be displayed as well as suction piping pressures, as indicated. The parameters displayed are determined by the system of the invention which utilizes the various sensors.

Various pressure sensors 452 sense pressure in the respective pump chambers associated with each of the pistons and pressure signals are transmitted to the processor. These pressure signals may indicate when valves are opening and closing. For example, if the pressure sensed in a pump chamber does not rise essentially instantly, after the piston for that chamber passes bottom dead center by 0 degrees to 10 degrees of crankshaft rotation, then it is indicated that the inlet or suction valve is delayed in closing or is leaking. In situations like this, the display may show that a discharge valve is not closed for 16.7 degrees of rotation after piston top dead center position. Accordingly, pressure changes, or the lack thereof, are sensed by cylinder chamber pressure sensors.

Software embedded in the computer 450 processor is operable to correlate the angle of rotation of the crankshaft with respect to pressure sensed in the respective cylinder chambers to determine any delay in pressure changes which could be attributable to delays in the respective suction or discharge valves reaching their fully seated and sealed positions. These delays can, of course, affect volumetric efficiency of the respective cylinder chambers and the overall volumetric efficiency of the pump. In this regard, total volumetric efficiency is determined by calculating the average volumetric efficiency based on the angular delay in chamber pressure increase or pressure decrease, as the case may be, with respect to the position of the pistons in the respective chambers.

The volumetric efficiency of the pump is a combination of normal pump timed events and the sealing condition of the piston seal and the inlet and discharge valves. Pump volumetric efficiency and component status is determined by determining the condition of the components and calculating the degree of fluid bypass. Pump volumetric efficiency (VE) is computed by performing a computational fluid material balance around each pump chamber.

Pump chamber pressures, as sensed by the sensors may be used to determine pump timing events that affect performance, such as volumetric efficiency, and chamber maximum and minimum pressures, as well as fluid compression delays. Still further, fluid pressures in the pump chambers may be sensed during a discharge stroke to determine, through variations in pressure, whether or not there is leakage of a piston packing or seal, such as the packing seal. Still further, maximum and minimum chamber fluid pressures may be used to determine fatigue limits for certain components of a pump, such as the fluid end housing, the valves and virtually any component that is subject to cyclic stresses induced by changes in pressure in the pump chambers and the pump discharge piping.

As mentioned previously, the computer 450 processor may be adapted with a suitable computer program to provide for determining pump volumetric efficiency which is the arithmetic average of the volumetric efficiency of the individual pump chambers as determined by the onset of pressure rise as a function of crankshaft position (delay in suction valve closing and seating) and the delay in pressure drop after a piston has reached top dead center (delay in discharge valve closing and seating).

Additional parameters which may be measured and calculated in accordance with the invention are the so-called delta volumes for the suction or inlet stabilizer and the discharge pulsation dampener. The delta volume is the volume of fluid that must be stored and then returned to the fluid flowstream to make the pump suction and discharge fluid flow rate substantially constant. This volume varies as certain pump operating parameters change. A significant increase in delta volume occurs when timing delays are introduced in the opening and closing of the suction and discharge valves. The delta volume is determined by applying actual angular degrees of rotation of the crankshaft with respect the suction and discharge valve closure delays to a mathematical model that integrates the difference between the actual fluid flow rate and the average flow rate.

Another parameter associated with determining component life for a pump, is pump hydraulic power output for each pump working cycle or 360 degrees of rotation of the crankshaft. Still further, pump component life cycles may be determined by using a multiple regression analysis to determine parameters which can project the actual lives of pump components. The factors which affect life of pump components are absolute maximum pressure, average maximum pressure, maximum pressure variation and frequency, pump speed, fluid temperature, fluid lubricity and fluid abrasivity.

As mentioned previously, pressure variation during fluid “compression” is an indication of the condition of a piston or plunger packing seal. This variation is defined as an absolute maximum deviation of actual pressure data from a linear value representative of the compression pressure and is an indication of the condition of seals. A leaking seal results in a longer compression cycle because part of the fluid being displaced is bypassing or leaking through the seal. A pump chamber “decompression” cycle is also shorter because, after the discharge valve completely closes and seals against its seat, part of the fluid to be decompressed is bypassing a plunger seal or packing. The difference in volume required to reach discharge operating pressure over a “compression” cycle for each pump chamber determines an average leakage rate. This leakage rate is adjusted for a leak rate at discharge operating pressures by calculating a leak velocity based on standard orifice plate pressure drop calculations.

Suction valve leak rate results in a longer decompression cycle because part of the fluid being displaced by the pressurizing element is returning to the pump inlet or suction fluid flowline. The difference in volume required to reach discharge operating pressure over a compression cycle determines an average leakage rate. This compression leak rate is then adjusted for a leak rate at discharge operating pressures by calculating a leak velocity based on standard orifice plate pressure drop calculations. The leak rate is then applied to the duration of the discharge valve open cycle. So-called pump intake or suction acceleration head response is an indicator of the suction piping configuration and operating conditions which meet the pump"s demand for fluid. This is defined as the elapsed time between the suction valve opening and the first chamber or suction piping or manifold pressure peak following the opening.

Still further, the system of the present invention is operable to determine fluid cavitation which usually results in high pressure “spikes” occurring in the pumping chamber during the suction stroke. Generally, the highest pressure spikes occur at the first pressure spike following the opening of a suction valve. Both minimum and maximum pressures are monitored to determine the extent and partial cause of cavitation.

The system is also operable to provide signals indicating valve design and operating conditions which can result in excessive peak pressures in the pumping chambers before the discharge valve opens, for example. These peaks or so-called overshoot pressures can result in premature pump component failure and excessive hydraulic forces in the discharge piping. For purposes of such analysis, the overshoot pressure is defined as peak chamber pressure minus the average discharge fluid pressure.

The system of the present invention is also operable to analyze operating conditions in the pump suction and discharge flow lines, such as in the piping. A normally operating multiplex power pump will induce pressure variations at both one and two times the crankshaft speed multiplied by the number of pump pistons. Flow induced pressure variation is defined as the sum of the peak-to-peak pressure resulting from these two frequencies. Also, acceleration induced pressure spikes are created when the pump valves open and close. Acceleration pressure variation for purposes of the methodology of the invention is defined as the total peak-to-peak pressure variation.

Hydraulic resonance occurs when a piping system has a hydraulic resonant frequency that is excited by forces induced by operation of a pump. Fluid hydraulic resonance is determined by analysis of the pressure waves created by the pump to determine how close the pressure response matches a true sine wave. The computer 450 is programmed to activate an alarm when the flow induced pressure variation exceeds a predetermined limit. Alternatively the computer 450 monitoring software can be programmed to trigger the reset relief valve that is operably connected and controlled by the processor.

Those skilled in the art will appreciate that the system, including pressure sensors together with the reset relief valve and its associated sensors provides information which may be used to analyze a substantial number of system operating conditions for a pump. The processor is adapted to provide a visual display 456 which may be displayed on the monitor, providing graphical display of pressure versus crankshaft position for each cylinder chamber and other parameters.

The system"s computer 450 controller, hard drive or other digital storage device and display 456 can be used for predictive analysis of the pump and component parts providing data to the operator to reveal conditions and maintenance related required tasks including but not limited to:Hidden Failure—A functional failure whose effects are not apparent to the operating crew under normal circumstances if the failure mode occurs on its own.

In conjunction with the pump, is a fully controlled and monitored pressure relief valve that is situated to set off with excessive mud flow pressures. Turning now to FIG. 5, there is a partially exploded cross sectional view of a reset relief valve showing chamber 500 and in fluid communication therewith and pressure sensor assembly 502 of the relief valve. A break is shown in FIG. 5 between the upper and lower portion of the valve where the pressure sensor assembly 502 is placed for ease of illustration. Transducer 506 is positioned in a sealed ring 504 that is in fluid communication with the hydraulic fluid that moves between chambers 500 and 512 during operation. On the bottom of ring 504 are ports 514 that permit fluid flow from chamber 500 into the ring whereby the transducer 506 can sense pressure in the fluid. Pressure is sensed by the transducer 506 which is capable of sensing the pressure on the fluid which in turn transmits that pressure reading to a gauge 516 and electronically to a computer control system. Gauge 516 may be analog or digital depending on the application. As pressure is being monitored and data points stored, the user is capable of controlling the valve and making sure its operation is within desired operating limits before during and after the valve is activated. Further, the computer parameters can be set for high pressures to shut down the pump before the reset valve is set off.

By storing data on a recurring basis, the operator can design the system with greater degrees of control and can analyze the data associated with an activation of the valve to better utilize the valve and other pumps in the system. Set screw 520 permits access to the system for bleeding off pressure. Other ports are positioned around the pressure sensor assembly for insertion of oil or other hydraulic fluid.

One of the advantages of having a relief valve that is computer monitored and controlled is that upon set-off the operator obtains valuable real time information about the pump just prior to and during set-off. The system is capable of storing data about the operation of the pump such as the time of set-off, the exact pressure that caused the relief valve to activate and speed and pressure associated with the pump.

The graphic display of the computer system may also show the discharge pressure parameters including discharge manifold pressure, total peak-to-peak pressure, flow induced peak-to-peak pressure, flow induced peak-to-peak pressure as a percent of average manifold pressure, the primary (largest) peak-to-peak pressure which is occurring at a particular frequency, the primary peak-to-peak pressure as a percent of average manifold pressure.

The processor may show pressure variation versus pump speed as determined by the system based on measuring chamber pressure and crankshaft position and speed.

The system may further display information showing crankshaft angle versus pump speed in strokes per minute showing discharge valve sealing delays in degrees of crankshaft rotation from piston top dead center. Suction valve sealing delays, from piston bottom dead center, may also be indicated.

The system of the invention may also be adapted to provide graphic displays such as a diagram of pump discharge pressure versus crankshaft angle showing the variation in pump discharge piping pressure, as well as the frequency and amplitude of pressure pulsations. Another display which may be provided by the system comprises a diagram of pump discharge piping pressure as measured by a pressure sensor versus pump speed in piston strokes per minute as calculated by the system. Still further, the system is operable to display fluid pressure conditions in the pump suction manifold, such as the manifold or piping. The system could also aid in de-synchronizing multiple pumps to decrease pulsation on the piping system.

A typical installation of a system for temporary or permanent performance monitoring and/or analysis requires that all of the pressure transducers be preferably on the horizontal center line of the pump piping or pump chambers, respectively, to minimize gas and sediment entrapment.

The system of the invention is also operable to determine pump piping hydraulic resonance and mechanical frequencies excited by one or more pumps connected thereto for both fixed and variable speed pumps. Preferably, a test procedure would involve instrumenting the pump, where plural pumps are used, that is furthest from the system discharge flowline or manifold. A vibration sensor, could be located at the position of the most noticeable piping vibration. The piping system should be configured for the desired flow path and all block valves to pumps not being operated should be open as though they were going to be operating. The instrumented pump or pumps should be started and run at maximum speed for fifteen minutes to allow stabilization of the system. A data acquisition system should then be operated to collect one minute of pumping system data. Alternatively, data may be continued to be collected while changing pump speed at increments of five strokes per minute every thirty seconds until minimum operating speed is reached. Data may be continued to be collected while changing suction or discharge pressures. The displays provided by the processor could be reviewed for pump operating problems as well as hydraulic and mechanical resonance. If a hydraulic resonant condition is observed, this may require the installation of wave blockers or orifice plates in the system piping.

The system is operable to provide displays comprising simulated three dimensional charts displaying peak-to-peak pressures occurring at respective frequencies for a given pump speed in strokes per minute. For a triplex pump, the normal excitation frequency is three and six times the pump speed. As pump speed increases, the excitation frequencies increase.

The software developed for the pump and valve is a method of acquiring multiple streams of analog data as real time occurrences and simultaneously displaying and storing them via digital interface using an interface system for later analysis. Using an application developed for this system, the control computer can monitor and record from 1 to 4 simultaneous analog pressure readings at a rate of 1 sample/sec per channel or faster. In addition a secondary capacity to monitor 4 digital only inputs is extant in the current system, with desired inputs undetermined. The data acquisition and its interaction and use with the software acts as a control system to the overal pump and its various operating compenents and attachments.

The computer system of a preferred embodiment may be of any of a variety of known systems with associated input and output devices, having adequate processing speed and storage capacity to analyze and process data associated with the operation of the pump. The system may also have wireless capability and preferably is capable of operating in temperature ranges of: 0 to 40° C.; Storage: −20 to 65° C., with Relative Humidity: 30% to 80%. It may use data processing of4 channels of 14-bit analog input

The present invention combines a computer controlled monitoring system for a pump and a reset relief valve with increased power characteristics and longevity by use of poly-fluorocarbon treated or infused materials and an integrally forged crank. By having the pump responsive to ambient conditions that are monitored real time, the reset valve, which is interconnected through sensors that sense operating conditions of the pump system, can send feedback to the motor to reduce pressure buildup before the valve is set off. Further, in certain conditions the reset valve can trip in advance of catastrophic failure of the pump through the computer control and feedback system of the present invention.

This present invention is directed to drilling wellbores in the earth, to systems for pumping drilling fluid (“mud”) for such operations, to mud pumping system modules with surge suppressing dampeners, and to methods of their use. DESCRIPTION OF THE RELATED

Known references disclose a wide variety of drilling systems, apparatuses, and methods including, but not limited to, the disclosures in U.S. Pat. Nos. 6,944,547; 6,918,453; 6,802,378; 6,050,348; 5,465,799; 4,995,465; 4,854,397; and 3,658,138, all incorporated fully herein for all purposes. Prior references disclose a wide variety of drilling fluid pumps (“mud pumps”) used in drilling operations and pump systems, for example, and not by way of limitation, those pumps and systems disclosed in U.S. Pat. Nos. 6,257,354; 4,295,366; 4,527,959; 5,616,009; 4,242,057; 4,676,724; 5,823,093; 5,960,700; 5,059,101; 5,253,987; in U.S. application Ser. No. 10/833,921 filed Apr. 28, 2004(all said U.S. references incorporated fully herein for all purposes). Known references disclose a variety of dampeners, accumulators, and surge suppressors; including, but not limited to, those disclosed in U.S. Pat. Nos. 4,299,253; 4,195,668; 2,757,689; 2,804,884; 3,674,053; 3,169,551; 3,674,053; 3,162,213; 2,380,866; 2,378,467; 2,397,248; 2,397,796; and 2,773,455—all incorporated fully herein for all purposes.

A drill bit carried at an end of a drillstring is rotated to form wellbores in the earth. Certain drillstrings include tubulars which may be drill pipe made of jointed sections or a continuous coiled tubing and a drilling assembly that has a drill bit at its bottom end. The drilling assembly is attached to the bottom end of the tubing or drillstring. In certain systems, to drill a wellbore, the drill bit is rotated (e.g., by a top drive, a power swivel, a rotary table system, or by a downhole mud motor carried by the drilling assembly). Drilling fluid, also referred to as “mud,” is pumped through the wellbore under pressure from a pit or container at the surface by a pumping system at the surface.

In certain known mud pump systems, suction and discharge modules have valves therein that selectively control fluid flow through the module in an intake (suction) mode in which piston apparatus creates a vacuum drawing drilling fluid into the module and in an output mode (Discharge) in which the piston apparatus creates pressure forcing drilling fluid out of the module. In the suction mode, a suction valve opens allowing drilling fluid into the module while a discharge valve remains closed. In the discharge mode, the pressure of the drilling fluid closes the suction valve and opens the discharge valve.

Both valves, the suction valve and the discharge valve, are subjected to the erosive and damaging effects of the flow of drilling fluid. The drilling fluid contains drilled cuttings and debris which can erode valve parts (e.g. seats, stems, valve members, seals, guide bushings, insert, liners, wear plates etc.). Also, mud pumps which can pump relatively hot drilling fluid at, e.g., 500 to 2000 gallons per minute, force the erosive drilling fluid against the valve parts at high velocities which add to the fluid"s damaging effects.

In many valves used in mud pump systems, a guide in the valve which is disposed across a flow path or guide fingers extending from a valve member into a valve seat guide a valve member so that valve member seats correctly and effectively against the valve seat. In many valves, the valve seat surface against which the valve member (or poppet) seats is, ideally, flat; and the surface of the valve member which sealingly abuts the flat seat surface of the valve seat is, correspondingly, and ideally, flat. A guide or guide fingers facilitates correct seating of the valve member"s flat seating surface against the valve seat"s flat seat surface. If either surface is not flat, or if one surface does not contact the other in a substantially parallel (flat surface to flat surface) manner, ineffective or inefficient valve operation may result.

In many known mud pump valves, the valves are opened and closed by mechanically creating a vacuum or fluid pressure increase in the valve that overcomes a spring to allow a valve member to move. The movement of the valve member is not controlled, i.e., it is subject to a surge of fluid under pressure. As fluid pressure builds up to move a valve member, a corresponding amount of fluid builds up adjacent the valve. when the pressure is high enough, a relatively large charge of fluid goes through the valve at high velocity. This surge of fluid can have deleterious effects on valve parts. BRIEF SUMMARY OF THE INVENTION

The present invention, in at least certain embodiments, discloses systems for pumping a drilling fluid mixture, the drilling fluid mixture containing drilling fluid and solids, the systems having: a pump apparatus; the pumping apparatus having a body with a pumping chamber, an inlet and an outlet; a suction valve in the body for selectively controlling flow of the drilling fluid mixture in through the inlet; a discharge valve in the body for selectively controlling flow of the drilling fluid mixture out through the outlet; and a dampener system according to the present invention in fluid communication with the pumping chamber.

Such a pump system according to the present invention, in one aspect, includes: a base; a housing connected to the base, the housing having an interior; a liner within the housing, the liner expandable in response to fluid pressure; a piston/cylinder apparatus in fluid communication with the housing; the piston/cylinder apparatus having a movable piston movable in response to fluid flowing from the housing to the piston/cylinder apparatus; a torsion apparatus movably connected to the base, the piston movable to contact and to move the torsion apparatus in response to fluid flowing from the housing to the piston/cylinder apparatus; and the torsion apparatus movable by the piston from a first static position to a second position to dampen pulsations of fluid into the pumping chamber.

In one aspect, a pumping system according to the present invention has a dampener system according to the present invention which includes: a housing, the housing having an interior; a deformable bladder within the housing, the deformable bladder in fluid communication with the pumping chamber; and the deformable bladder deformable in response to pressure variation in the pumping chamber.

The present invention discloses, in certain aspects, dampeners for drilling fluid pumping systems which suppress and/or eliminate the damaging effects of undesirable pulsations or surges of drilling fluid passing through the systems. In certain aspects, the dampener has a liner with liquid therein which expands and contracts in response to the pressure of drilling fluid passing through a pumping system.

The present invention discloses, in certain aspects, dampeners for drilling fluid pumping systems in which the dampener has a liner with liquid therein which expands and contracts in response to the pressure of drilling fluid passing through a pumping system. In certain aspects, a dampener according to the present invention has a torsion apparatus that absorbs and then releases energy to facilitate the dampening of drilling fluid surges. In other aspects, a dampener system according to the present invention has an inflatable bladder surrounded by an expandable spring member, both the bladder and the spring member responsive to drilling fluid surges to suppress deleterious effects of such surges.

The present invention discloses, in certain aspects, modules for a drilling fluid pumping system which include a dampener for suppressing and/or eliminating the damaging effects of undesirable pulsations or surges of drilling fluid passing through the modules. In certain aspects, the dampener is within a block of the module that also contains suction and discharge valve assemblies within a module block.

The present invention discloses, in certain aspects, a drilling fluid pumping system, also known as a mud pump system, for pumping drilling fluid or mud used in wellbore operations which has pumping modules with valves that have non-flat seating surfaces. In certain aspects, such valves have a valve member or poppet that is movable with multiple degrees of freedom in any of which effective seating of the valve member against a valve seat is achieved. In particular aspects of such a valve, dual sealing is achieved by sealing of a valve member against both a valve seat and against a seal disposed in a valve seat.

In certain particular aspects of a mud pump system according to the present invention, a mud pump valve has a tapered spring biased against a valve member which enhances the free seating movement of a valve member.

The present invention discloses, in certain aspects, valves for a system for pumping a drilling fluid mixture, the drilling fluid mixture containing drilling fluid and solids, the valves having: a seat with a valve seat surface; a valve member with a member surface, part of the valve member movable to seat the member surface against the valve seat surface to prevent the flow of the drilling fluid mixture past the valve seat; a cartridge stem positioned with respect to the valve member, and a valve actuator within the cartridge stem for selectively moving the valve member. In certain aspects, the present invention discloses a system f