strokes per minute mud pump free sample

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Rig pump output, normally in volume per stroke, of mud pumps on the rig is  one of important figures that we really need to know because we will use pump out put figures to calculate many parameters such as bottom up strokes,  wash out depth, tracking drilling fluid, etc. In this post, you will learn how to calculate pump out put for triplex pump and duplex pump in bothOilfield and Metric Unit.

Oil and Gas drilling process - Pupm output for Triplex and Duplex pumpsTriplex Pump Formula 1 PO, bbl/stk = 0.000243 x ( in) E.xample: Determine the pump output, bbl/stk, at 100% efficiency for a 7" by 12". triplex pump: PO @ 100%,= 0.000243 x 7 x12 PO @ 100% = 0.142884bbl/stk Adjust the pump output for 95% efficiency: Decimal equivalent = 95 + 100 = 0.95 PO @ 95% = 0.142884bbl/stk x 0.95 PO @ 95% = 0.13574bbl/stk Formula 2 PO, gpm = [3(D x 0.7854)S]0.00411 x SPM where D = liner diameter, in. S = stroke length, in. SPM = strokes per minute Determine the pump output, gpm, for a 7" by 12". triplex pump at 80 strokes per minute: PO, gpm = [3(7 x 0.7854) 1210.00411 x 80 PO, gpm = 1385.4456 x 0.00411 x 80 PO = 455.5 gpm

Example:Duplex Pump Formula 1 0.000324 x (liner diameter, in) x ( stroke lengh, in) = ________ bbl/stk -0.000162 x (rod diameter, in) x ( stroke lengh, in) = ________ bbl/stk Pump out put @ 100% eff = ________bbl/stk Example: Determine the output, bbl/stk, of a 5 1/2" by 14" duplex pump at 100% efficiency. Rod diameter = 2.0": 0.000324 x 5.5 x 14 = 0.137214bbl/stk -0.000162 x 2.0 x 14 = 0.009072bbl/stk Pump output @ 100% eff. = 0.128142bbl/stk Adjust pump output for 85% efficiency: Decimal equivalent = 85 100 = 0.85 PO@85%)= 0.128142bbl/stk x 0.85 PO@ 85% = 0.10892bbl/stk Formula 2

PO. bbl/stk = 0.000162 x S[2(D) - d] where S = stroke length, in. D = liner diameter, in. d = rod diameter, in. Example: Determine the output, bbl/stk, of a 5 1/2". by 14". duplex pump @ 100% efficiency. Rod diameter = 2.0in.: PO@100%=0.000162 x 14 x [ 2 (5.5) - 2 ] PO @ 100%)= 0.000162 x 14 x 56.5 PO@ 100%)= 0.128142bbl/stk Adjust pump output for 85% efficiency: PO@85%,= 0.128142bb/stkx 0.85 PO@8.5%= 0.10892bbl/stk Metric calculation Pump output, liter/min = pump output. liter/stk x pump speed, spm. S.I. units calculation Pump output, m/min = pump output, liter/stk x pump speed, spm. Mud Pumps Mud pumps drive the mud around the drilling system. Depending on liner size availability they can be set up to provide high pressure and low flow rate, or low pressure and high flow rate. Analysis of the application and running the Drill Bits hydraulics program will indicate which liners to recommend. Finding the specification of the mud pumps allows flow rate to be calculated from pump stroke rate, SPM. Information requiredo Pump manufacturer o Number of pumps o Liner size and gallons per revolution Weight As a drill bit cutting structure wears more weight will be required to achieve the same RoP in a homogenous formation. PDC wear flats, worn inserts and worn milled tooth teeth will make the bit drill less efficiently. Increase weight in increments of 2,000lbs approx. In general, weight should be applied before excessive rotary speed so that the cutting structure maintains a significant depth of cut to stabilise the bit and prevent whirl. If downhole weight measurements are available they can be used in combination with surface measurements to gain a more accurate representation of what is happening in the well bore.

The 2,200-hp mud pump for offshore applications is a single-acting reciprocating triplex mud pump designed for high fluid flow rates, even at low operating speeds, and with a long stroke design. These features reduce the number of load reversals in critical components and increase the life of fluid end parts.

The pump’s critical components are strategically placed to make maintenance and inspection far easier and safer. The two-piece, quick-release piston rod lets you remove the piston without disturbing the liner, minimizing downtime when you’re replacing fluid parts.

Kverneland, Hege, Kyllingstad, Åge, and Magne Moe. "Development and Performance Testing of the Hex Mud Pump." Paper presented at the SPE/IADC Drilling Conference, Amsterdam, Netherlands, February 2003. doi: https://doi.org/10.2118/79831-MS

NOV 12-P-160 Mud Pump is rated at 1600 input horsepower (1193 kw) at 120 strokes per minute, with a 12-inch (304.8 mm) stroke. Multiple liner sizes allow pressures and volumes to handle circulation requirements in deep drilling applications.

Flexibility: Compact engineering provides higher efficiency in less space. The NOV 12-P-160 Triplex Mud Pump light weight and flexible design make it easily adaptable to a variety of rig configurations. This provides flexibility as drilling requirements and conditions change.

Fluid End Modules: NOV offers a choice of fluid end modules and valve covers for every P Series pump model to select the fluid end module that exactly matches drilling requirements. All pump models can be equipped with either the standard or premium forged, two-piece interchangeable fluid modules

Our pump stroke counter systems (CPS101 Series) measure the stroke rate and number of strokes on mud pumps. The oilfield pump stroke system is user-friendly and reliable and is configurable to measure up to three mud pumps at once. Our digital pump stroke counter systems are manufactured here in the U.S. by Crown Oilfield Instrumentation, and Crown’s Pump Stroke Counter provides easy monitoring of strokes per minute on multiple mud pumps. Each mud pumps’s stroke rate can be selected individually and the display is updated regularly for accurate monitoring. LCD displays indicate both pumps strokes per minute and the total number of strokes. Located at the bottom of the panel, push buttons provide easy operation and reseting of each pump. When you need to accurately monitor and maintain the amount of mud being pumped, you can trust Crown’s oilfield stroke counters.

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.

Yet another advantage of the invention is to provide a single piece integrally forged and balanced crank assembly for greater strength during operation.

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.

Lubrication Task—The periodic application of a lubricant to items that require lubrication for proper operation or to prevent premature functional failures.

On Condition Task—A periodic or continuous inspection designed to detect a potential failure condition and allow correction prior to functional failure.

Preventive Maintenance (PM)—Actions performed prior to functional failure (multiple failures or demand requirements for hidden failures) to achieve the desired level of safety and reliability for an item.

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 software application employed preferably is a data acquisition and storage system with a 90 day continuous storage capacity. The computer software permits real time evaluation of ambient pressure in the system and control of the valve according to the user"s preferences. Due to the recurrent sensing of the system, when the valve is set off, data has been stored for later analysis of the cause of the pressure spike that may have set off the valve.

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 invention relates generally to piston pumps for the water well drilling industry, and more particularly to a hydraulic cylinder powered double acting duplex piston pump.

Double acting duplex piston pumps are well known and have been used in the water well drilling industry for many years. Typically they employ a crankshaft and flywheel driven in various ways, a reciprocating engine or a hydraulic motor being examples. Typically, they are heavy units with a large component of cast iron. Today"s well drilling trucks carry lengths of drilling pipe, as well as derricks, motors, pumps of various kinds, and the “mud” pump. The current double acting duplex piston pumps with crankshaft and flywheel, being very heavy, contribute significantly to the weight and space requirements of the truck. They impact the ability of a truck to meet federal highway weight restrictions. Also, the mechanical crank throw design imparts a variable speed to the mud pump piston. In such designs, the piston is either accelerating or decelerating during a large part of its design stroke. So the piston operates at its full design capacity during only a portion of the stroke. Therefore, it is an object of the present invention to provide a duplex piston pump useful as a mud pump on a water well drilling machine, but without a motor, crankshaft, flywheel, gearing, and/or belts, for a significant weight reduction.

Described briefly, according one embodiment of the present invention, a mud pump is provided with two working cylinders for pumping mud, and two sets of double-acting hydraulic driving cylinders. One set of two driving cylinders has the piston of each connected to the piston of one double-acting mud pump cylinder. The other set of two driving cylinders has the piston of each connected to the piston of the other double-acting mud pump cylinder. The connection of a set of driving cylinder pistons to the mud pump piston is through a member which allows side-by-side, or over and under parallel arrangement of the driving cylinders and mud pump cylinders, so the overall length is minimal. An electro-hydraulic control system is provided to coordinate the action of the pump driving cylinders with the mud pumping cylinders for contributing to steady flow of mud from the mud pump.

FIG. 2 is a schematic elevational view of one of the two pump assemblies of the duplex piston pump according to one embodiment of the present invention.

FIG. 3 is a schematic diagram of a system according to the illustrated embodiment and showing both of the duplex mud pump cylinders, with one of two hydraulic driving cylinders for each of the mud pump cylinders, and including an organization of hydraulic flow divider, rod position sensing, proximity-type electrical switches and associated electrical relays for solenoid-operated hydraulic fluid directing spool valves associated with the hydraulic cylinders.

FIG. 6 is an example of a flow chart relating mud pump speed to mud pump output volume capacity and hydraulic driving oil volume requirement for a pump according to the illustrated embodiment of the present invention.

FIG. 7 is a diagram showing theoretical pump driving cylinder piston performance of the two sets of mud pump driving cylinders operating according to the illustrated embodiment of the present invention.

FIG. 1 shows, schematically, a normal environment in which the mud is pumped by duplex piston pump 5 of the present invention from the chip separation tank 6 through the pump and the discharge the line 7 into the top of drill pipe 8 and down in the drill pipe and out into the well casing at the drill bit 10. The mud flows upward through the casing and back into the separation tank 6. The pump itself includes two mud pumping cylinders 1 and 2 fixed relative to a base 3 (FIGS. 5A and 5B) by mounting to four housings 4L, 4F, 4R and 4B which are fixed to the base.

As suggested above, according to the present invention, an all-hydraulic drive for the two mud-pumping pistons in cylinders 1 and 2 is achieved. For that purpose, and referring to FIG. 3, a variable volume hydraulic pump P is used. It can be set, for example, at a rate of 36 gallons per minute at 1,000 psi. A motor M can be used to drive such a pump, and such pumps and drives for them are known in the art and readily available. The pump output is fed to a flow divider D. This is not merely a device to split the flow. Instead, it has a piston inside which will shift in either direction to the extent necessary as it tries to be sure that the exact same volume flow rate is delivered at both output ports of the flow divider. An example of such a flow divider is Model MH2FA by Rexroth Worldwide Hydraulics.

Before proceeding further with this description, it is important to understand that each half of the duplex piston pump includes a double-acting mud pump cylinder and piston assembly. FIG. 2 shows one of the halves including mud pump cylinder 1. The other half, including a mud pump cylinder and associated hydraulic driving cylinders is identical. FIG. 5A shows, schematically, a rod end view of the duplex piston mud pump 5 including the mud pump cylinders 1 and 2 and associated driving cylinders of both halves.

The mud pump cylinders and their associated driving cylinders may be fixed relative to each other and mounted to the base 3 by any suitable means. The schematics of FIGS. 5A and 5B show, as an example, the four valve housings 4F, 4L,4B and 4R fixed to base 3. Except for right and left ports, each of these housings is the same as the others, and includes therein a chamber such as 41C having an opening 41A communicating with a port of a mud pump cylinder such as cylinder 1 or 2. Each chamber 41C also has two other openings. One of them is fitted with a one-way, spring loaded outlet valve such as 16 to enable mud to move from chamber 41C into the upper end of housing 4L and out through port 41D into the discharge plenum 7A. The other opening 41E of chamber 41C is fitted with a spring loaded inlet valve such as 14 communicating with the lower end of housing 4L and enabling mud entering suction inlet 6A of the mud suction manifold 6B to enter through port 41E in housing wall and into chamber 41C. The mud pump cylinders are attached to their respective valve housings in conventional manner with the ports of the cylinders communicating with the respective chambers of the housings.

Referring now to FIGS. 2 and 5A and 5B together, FIG. 2 shows schematically mud pump cylinder 1 and inlet and outlet valves 11 and 12, respectively, associated with the pump port at one end of cylinder 1, and inlet and outlet valves 14 and 16, respectively, associated with the pump port at the opposite end of cylinder 1, the rod-end of the cylinder. In practice since there would normally be only one port at each end of the cylinder, the valves would be in the housings 4L and 4F for cylinder 1. There is a packing gland 17 at the rod-end of the cylinder. In FIG. 5B the top of housing 4L is cut away to show discharge valve 16. Suction or inlet valve 14 below chamber 41C is shown larger in dashed lines to be able to see it in FIG. 5B, as it is the bottom of chamber 41C, FIG. 5A.

According to one feature of this invention, there are two hydraulic driving cylinder/piston assemblies 19 and 21 arranged in a way to drive the mud pump piston 1P in cylinder 1. As suggested above, cylinders 1, 19, and 21 are all connected relative to each other by some suitable means (brackets and/or clamps, for example) so that they are longitudinally immovable relative to one another. This is represented schematically for both sets of mud pump and driving cylinders at the valve housing 4L, 4R and associated inlet and outlet and base in FIGS. 5A and 5B, to which all of the six cylinders of both halves of the duplex pump are rigidly, but removably connected. Each of the two driving cylinder assemblies for piston 1P has a piston rod such as 19R and 21R bolted to a rod connector plate 22. Each rod may extend through the piston and exit the driving cylinder at the end opposite the plate-connected rod end, as indicated at 19S and 21S. The respective pistons 19P and 21P are affixed and sealed to the rod in a suitable way and may be of any suitable construction. Of course, the same effect may be achieved using separate colinear rods secured to opposite faces of the pistons. The piston rods 19R and 21R of the driving cylinders for mud pump cylinder 1, and rod 1R of the mud pump cylinder 1 itself, are bolted to the rod connector plate 22. Each of the rods 19R, 19S and 21R, 21S is supported at opposite ends of the respective cylinder by bearings and seals at 19B and 21B, respectively. Using double rod cylinders provides equal working area on the two sides of the piston, enabling equal oil flow and thrust capacity in both directions of piston travel. One or the other of the driving cylinder rods for mud pump cylinder 1 is associated with a set of proximity sensing switches A-1, B-1, C-1, and D-1 to operate a relay to shift a hydraulic solenoid valve spool to control hydraulic fluid to and from the set of two driving cylinders 19 and 21 to drive the mud pump 1. The same kind of arrangement is provided for mud pump cylinder 2. For each set, the driving cylinder which has the associated proximity switches, may be referred to hereinafter from time-to-time, as the control cylinder.

In this particular arrangement, only as an example of cylinder and rod size, the mud pump cylinder may be six inches in diameter with a twelve inch stroke, using the pistons of two driving cylinders of two inch diameter each to drive the one mud pump piston. A significant advantage can be achieved by making the rods of the driving cylinders larger in diameter (1.375 inches, for example) than that of the mud pump cylinder rod (1.25 inches, for example). It enables use of larger and longer wearing bearings in the driving cylinders, and enables the use of a relatively small piston rod and packing gland 17 in the mud pump cylinder, thus minimizing exposure to wear of the packing gland. The combination of the large diameter rods in the driving cylinders, fixed to a rigid rod connector plate 22 to which the mud pump cylinder rod is bolted, contributes to a very rigid structure. It avoids the necessity of a very long arrangement and long piston rod spans which would be necessary if the mud pump cylinder was driven by a single piston in a hydraulic cylinder on the same longitudinal axis. That would require a more complicated bearing arrangement to support the mud pump cylinder rod. In the present arrangement, the cylinder rod bearings are relatively close to the mud pump packing gland, helping extend the life of the gland by minimizing radial working and resulting loading of the mud pump rod on the packing gland. Also, with the present arrangement, the driving cylinder rods are in tension when the mud pump rod is in compression, which reduces the bending moment.

The proximity sensor switches A-1 through D-1 are responsive to movement of an actuator such as flange 19F on rod 19R, 19S. These switches may be normally-closed or normally-open switches as a matter of convenience in the construction of the circuitry. It should be understood, of course, that the other half of the duplex double acting pump assembly which includes cylinder 2, has driving cylinders such as 29 and 31 associated with it, and proximity switches associated with the piston rod of one (29, for example) of those driving cylinders, (shown in FIGS. 3 and 5A) in the same manner as for the assembly shown in FIG. 2. The position of the piston in cylinder 1 is preferably in a different location and/or it is moving in a different direction, from that of the piston in cylinder 2.

Referring now to FIG. 3, mud pump cylinder 1 and mud pump cylinder 2 are shown schematically, as is one cylinder of each set of two driving cylinders for each of the two mud pump cylinders 1 and 2, respectively. Since the mud pump cylinders are virtually identical and the two driving cylinders of the set for each of the mud pump cylinders are virtually identical, a description of one driving cylinder and associated controls will suffice for both.

The output from flow divider D enters the center input port of a two-position, solenoid-actuated, spring-return hydraulic valve V-1. This valve is electrically coupled to a relay switch R-1 which is bi-stable and electrically coupled to the proximity sensor switch A-1. An example of a suitable relay is No. 700-HJD32Z12 by Allen-Bradley. It is a DPDT latching relay. One switched position of this relay switch R-1 causes the solenoid to be energized to open the valve and supply pressurized oil from valve V-1 through line L-1 to the one end of cylinder 19 and likewise cylinder 21 of FIGS. 2 and 5 to drive the pistons in the direction of the arrow 23. This occurs in both driving cylinders 19 and 21, so rod 13R, being mechanically fixed to the two driving cylinder rods 19R and 21R by rod connector plate 22, is likewise driven in the direction of the arrow 23. When relay R-1 is reset by a signal from another proximity switch which can be recognized upon study of FIGS. 4A through 4D, it de-energizes the solenoid for valve V-1, enabling the spring therein to return the solenoid to position where the supply to the cylinder 19 is through line L-2, to reverse the direction of the piston in that cylinder and its companion driving cylinder, thus reversing the direction of the mud pump piston 1P. Whichever side of the piston is not pressurized at any time is enabled to dump through the valve V-1 to sump S-1. Essentially the same arrangement exists for control and drive of mud pump cylinder 2. In this instance, the proximity switches are designated A-2, B-2, C-2 and D-2. The relay switch is R-2 and the control valve is V-2 operated by a solenoid. It should be mentioned at this point, however, that while the pistons in the driving cylinders for one of the mud pump cylinders are located in the same relationship to each other as the mud pump cylinder with which they are associated, they are typically out of phase with respect to the mud pump piston and pistons of associated drive cylinders of the other mud pump cylinder. This is intentional in an effort to be sure that the flow out of the mud pump assembly 5 is as stable and constant as possible. That is the goal to which the organization of the proximity switches and associated relay switches are directed. Also with reference to FIGS. 2 and 3, it should be mentioned that the supply lines L-1 and L-2 to cylinder 19 are larger than the lines from cylinder 19 to cylinder 21. This is because the lines from valve V-1 go directly to only one of the two driving cylinders and from that point, are directed to the other driving cylinder. Thus, the supply lines L-1 and L-2 must be large enough to drive both driving cylinders 19 and 21 with essentially equal pressure and volume capacity. This is shown schematically in FIG. 2 with lines B-1 and B-2 from cylinder 19 to cylinder 21.

Referring now to FIGS. 4A through 4D, along with FIG. 3, FIG. 4A is a simplified portion of FIG. 3. It includes a driving cylinder 19 for mud pump cylinder 1, and driving cylinder 29 for mud pump cylinder 2. It also shows the proximity switches A-1 and B-1 associated with the piston rod portion 19S of cylinder 19. Similarly, proximity switches A-2 and B-2 associated with the piston rod 29S of driving cylinder 29, are shown. The position of the rod 19S relative to rod 29S is only for purposes of example, as it is not expected that the pistons of cylinders 1 and 2 will ever be positioned at the same longitudinal location relative to each other unless they are passing as one goes in one direction and the other goes in the other direction. But the purposes of the proximity switches A-1 and B-1 is to limit the travel of the piston in the two directions. Thus, either the switch A-1 or the switch B-1 can set or reset relay R-1 to cause the valve V-1 to shift and switch the high pressure from valve V-1 to either line L-1 to move the driving piston and thereby the mud pump piston 1P to the right,