installing element in mud pump pulsation dampner manufacturer

A properly serviced pulsation dampener is critical for your mud pumps’ efficiency, safety, and performance. Unfortunately, there aren’t many resources available to educate personnel on executing safe and effective servicing procedures. Please review the following steps with your personnel for safe pulsation dampener maintenance.

Should you or your personnel have any questions regarding pulsation dampener maintenance, please don’t hesitate to ask. Sigma is more than happy to help you to ensure safe and proper care is being completed on your pulsation dampening equipment.

Pulsation dampeners (also called pulsation dampers) are used for stabilizing the flow and the pressure in circuits with volumetric or dosing pumps. They are used in a wide range of applications.

In every pulsation dampener there is a separator element between the gas it is charged with and the liquid of the circuit; its basic function being to avoid the leaking of the gas into the circuit. This separator element is basically made of two kinds of materials: Rubber (NBR, EPDM, FPM, butyl, silicone, etc…) or a thermoplastic material (normally PTFE); although it can also be made in stainless steel.

When a rubber separator element is used, the dampener is called bladder type. If the material is PTFE, we refer to membrane type and bellows type dampeners, depending on the shape of the separator element.

Choosing between different types of dampener depends on characteristics of the circuit like working pressure, temperature and chemical compatibility between the liquid and the material of the separator.

All our pulsation dampenersare made according to the European PED97/23/CE pressure vessels regulations, and their design meets the AD-2000 and ASME VIII Div.1 & 8 codes requirements (“U” stamp pending).

We can supply all of our dampeners with different circuit connection gauges as well as fitted with whatever flange, either screwed on, welded or integrated, to suit the customer’s needs.

Reciprocating pumps emit pulsations into the attached piping systems causing potentially dangerous unbalanced shaking forces, poor valve dynamics, high maintenance costs, and reduced flow. It is well known that gas charged elastomeric bladder or diaphragm pulsation dampeners can be used to effectively reduce the risks of these pulsation induced problems. Gas charged dampeners are relatively compact in size, and are easy to install when compared with maintenance free liquid filled options which tend to be large and more expensive to fabricate.

The potential for the dampener itself to become a vibration problem due to its branch connection being a heavy cantilevered mass having low mechanical natural frequencies.

These drawbacks are common issues for many operators. In a recent brownfield upgrade pump installation, Shell requested that the new pumps include a monitoring system to indicate when the gas charged dampener has failed. Beta Machinery (Beta) were contracted to include this study with the pulsation analysis being performed as per API 674. The hypothesis of the monitoring study was simple: when the dampener fails pulsations increase, an alarm is then triggered indicating a dampener failure.

The test of this hypothesis is presented in the following two-part case study of a brownfield upgrade. In the first part, we will provide the background of the oil company’s experience and the pulsation system model piping design. In next month’s installment, we will examine the complications that arise in a real installation and the challenges to overcome with variable speed pump operation.

Shell’s experience includes numerous reciprocating pumps for glycol and hydrocarbon condensate duties on both manned and unmanned offshore platforms. A majority of the older installations have reciprocating pumps equipped with bladder type pulsation dampeners. There have been numerous technical integrity issues related with the pulsation dampener losing pre-charge frequently, cases with the bladder material rupturing prematurely, and many cases of a notable increase in the vibration of the pump pipework leading to a failure and loss of containment.

With no form of indication available to ascertain the pre-charge pressure, it is difficult if not impossible to determine if the bladder, and/or bladder pressure is intact and holding. The only way to tell if the pulsation dampener is not performing as intended and has lost pre-charge is visual inspection of pipework. When the dampener has lost its pre-charge, the pipe work rattling and high vibration indicates some flaw in the working of the dampener. Under these circumstances the equipment has to be shutdown to manually inspect the dampener, as this has a direct impact on the reliability and downtime of the equipment.

The reciprocating pumps at the Shell Operating Unit facility are not equipped with on-line condition monitoring and the maintenance philosophy revolves around periodic off-line vibration condition monitoring typically carried out using a hand held instrument at three monthly intervals. For unmanned platforms the increase in vibrations are detected when the pump pipework vibration has increased substantially due to loss of pre-charge or the bladder giving away, presenting a high risk threat for operations.

The bladder in the dampener is very sensitive to changes in gas volume. As the line pressure changes, the volume of gas in the bladder changes, thus altering the system performance. The effectiveness of the pulsation dampeners can be reduced by bladder stiffness, bladder permeability, restriction of bladder expansion and contraction by dampener internals and degradation of performance due to large variations in line pressure, making the prediction of dampener performance complicated.

Technical integrity issues with the bladder material rupturing prematurely also surface when technicians charge the pulsation dampeners too quickly, leading to low temperatures at the elastomer interface and cracks propagating through the material. This has the effect of ultimately causing premature failure of the bladder material within a few charges.

Loss of containment through excessive pipe vibration and dampener failure, and the resulting impact on reliability statistics provided the motivation to address some of the issues described above. An opportunity to monitor the technical integrity of pulsation dampeners arose through brownfield project of F13 condensate transfer pumps (Triplex –double diaphragm type) to be installed on the existing E11 PB platform.

In order to address the pulsation dampener integrity issues, a change in the type of dampener from bladder type to the acoustic liquid filled type was considered. This approach was successful only on the greenfield projects. For the brownfield applications this approach met with a limited success considering the fact that space and cost of the liquid filled dampener is much higher than an equivalent gas filled dampener.

The existing piping arrangement is often congested, particularly for offshore facilities, hence the option of installing a liquid filled dampener is effectively ruled out. Moreover most of the pumps within the facility are slow speed with low frequency. Thus this approach was not effective in resolving the brownfield installations.

Metallic bellows type dampeners were next considered for cases where the bladder failed prematurely or failed to hold the pre-charge with the elastomeric element slipping off. This approach did provide some amount of success especially with the high temperature glycol service where the fluid temperature is around 248 degrees Fahrenheit (120 degrees Celsius). The high temperature for the glycol service led to the premature failures of the bladder elastomeric element and cases where the bladder lost its pre-charge within a few days of operation.

Manufacturers of pulsation dampeners were requested to provide some form of indication on the pulsation dampener for ascertaining the pre-charge of nitrogen. Some of the pulsation dampeners complied with this requirement by providing a T-connection with one end of the T connected to a pressure gauge. However, such indication is purely local and is of value only at manned platforms. For the unmanned platforms the issue of pre-charge pressure uncertainty remained.

With appropriate signal conditioning, accelerometers could be used to measure the low frequency and high frequency vibration. Accelerometers mounted on the pump manifolds serve this purpose. The acceleration data can then be used to identify issues with plungers/control rods or valves. However, vibration on the pump piping is generally due to the unbalanced forces in the piping from pulsation, as such measuring vibration would be an indirect method of gauging pulsation. So why not measure pulsation directly?

Based on the limited success with the above approaches, Beta was engaged to work with a manufacturer to provide a solution for triplex pumps to be installed on an unmanned facility. The objective of this exercise was to design an effective and well controlled pulsation system, and at the same time develop a monitoring system capable of detecting the condition of pulsation dampeners and alert operations for any abnormal behavior in the system. It is also to provide some form of assurance to predict the behavior of the dampeners under dynamic variations in the system pressure. F13 condensate transfer pumps were the subject of this approach.

In the second part of this analysis, we will illustrate that the difficulty and risk in maintaining a safe pump system with these dampeners is magnified in offshore and unmanned equipment operations.

Until better options appear on the market to address these important issues, pulsation monitoring is a technique that can be used to monitor dampener integrity. ■

Jordan Grose is the manager of pump systems for Beta Machinery Analysis. Ravindra Pai is a senior rotating engineer at Shell. For more information, visit www.betamachinery.com. This data was presented in an altered form at Dusseldorf’s International Rotating Equipment Conference in 2012.

Positive displacement pumps effectively pump fluid at a constant average flow rate. However, because the individual pumping elements of these pumps discharge discrete quantities of fluid, the instantaneous flow rate varies in a cyclic fashion.

Pulsations are observed in the system as pressure spikes. In the positive displacement pump family, single-shoe peristaltic pumps generally create the largest pulse, followed by two-shoe peristaltic pumps. Triplex and quintuplex pumps have smooth output curves because of piston overlap. Gear pumps can have extremely small pulses, but pulsations still exist. This pulsating flow can cause operational problems and shorten equipment’s service life.

To alleviate the problem, pulsation dampeners can be added to the pumping system to absorb pressure spikes and smooth fluid flow. Figure 1 shows the undampened pressure spikes from a triplex pump in green. The dampened pressure curve from the same pump with the same system settings are indicated in blue. Six pulses per revolution occur instead of the expected three. This is a result of piston overlap.

The most common type of pulsation dampener is a hydro-pneumatic pressure vessel containing compressed air or nitrogen and a bladder—or bellows—that separate the process fluid from the gas charge. To maximize the dampening effect, pulsation dampeners should be installed as close as possible to the pump discharge with a gas charge that is slightly below the normal system pressure. More important, pulsation dampeners must be properly sized for the system.

A dampener that is undersized cannot adequately compensate for pressure and flow fluctuations. An oversized dampener will act as an accumulator, storing too much fluid. This will cause slow stabilization and a delayed response to system changes. The first step in sizing a dampener is to quantitatively define the acceptable performance.

The specific requirements of the application and the components that make up the system are all factors that need to be considered. Once an acceptable pressure variation is defined, the unit size required for the desired performance should be determined. Engineers and designers are interested in making accurate predictions. Avoiding a problem is better than finding a way to fix it.

Sizing pulsation dampeners is straightforward. However, calculating the system pressure fluctuations is more complex. Fluid discharge rates from pumps are difficult to mathematically model. For example, in Figure 1, the spikes are not even. Theoretically, they should be equal. Mathematical models must be physically tested to verify their accuracy.

Pumps with multiple heads and higher pulse frequencies can make the calculations more difficult. The distance from one output port to the next is generally not constant. This creates a shift in the piston overlap with intermittent larger and smaller pulses. Calculating the magnitude or frequency of noise pulses that can develop or resonate in a system is difficult.

Piping arrangement—such as bends, reducers and valves—combined with the opening and closing of pump discharge check valves can create noise in the fluid called pressure pulses. Because many variables must be considered, each pump type should be tested with and without a dampener. The pressure curve data can be recorded and used to find the pump’s formula constant. This constant can be used in future calculations. As long as other pump models are similar to the test unit, accurately predicting the magnitude of line pressure variation with a given size dampener is possible.

The pressure in a piping system will rise sharply when a volume of fluid is added to the line. It accelerates the mass of the fluid in the piping system. This is acceleration head, and it needs to be minimized with a dampener. The effect and its impact must be considered on both the inlets and outlets of positive displacement pumps. On the inlet side, cavitation and partial filling of pump cavities can damage pump components and make the pump much louder than normal.

Bourdon tube gauges require time to equalize and can undershoot and overshoot the actual pressure depending on the magnitude and frequency of the pressure pulse. Even if the gauge could read accurately, reading a quickly moving dial is difficult. Electronically measured and recorded data can determine how the system is operating.

System noise must be considered when taking measurements because it can give higher-than-expected results. Noise in the pumping liquid can generally be ignored, but in some situations, system noise needs to be controlled. Noise can cause pressure relief valves to leak, damage sensitive components and create occupational safety hazards. Dampeners typically reduce noise, and some are specifically designed for this purpose.

Several different styles of dampeners are available, and each has advantages and disadvantages. This article focuses on reducing the pressure pulses caused by pulsing flow. The principles and the method for calculating the appropriate size dampener for this application are the same for most dampeners.

A dampener absorbs a fluid pulse and then allows the fluid to flow back into the system between pulses. Most dampeners use a gas charge that is set slightly below the normal system pressure and is compressed by the pulse of fluid. The gas then expands when fluid is released.

In this formula, n is a constant that is specific to the gas being used. For example, for air at room temperature, 
n ≅ 1.4, and for nitrogen, n ≅ 1.399.

Some heat transfer almost always occurs. The process is rarely slow enough for the gas temperature to equalize, so the actual answer will be between these two calculations. In most cases, the fluctuations are fast enough that the actual value is significantly closer to the isentropic formula. The isentropic formula gives the most conservative result. Therefore, it is the more accurate formula in most cases.

In actual practice, either formula would probably work if the pressure fluctuations are small relative to the system pressure. The pump constant that is developed would cover the inaccuracies in the formula as long as the pressure variations are similar. In this article, the isentropic formula is used.

To determine the pump constant, the volume from a single pulse of the pump must first be determined. Then an initial estimate of dampener size is made, and the corresponding value of dampener volume is applied. The amount of gas in the dampener will be less than the total dampener volume, which needs to be factored into the calculation. A typical range of 80 to 90 percent of the dampener volume should be gas if the dampener is properly charged. These give an initial gas volume:

The constant reduces the pulse volume to account for flow leaving the dampener while the pulse is entering. It also accounts for piston overlap, which changes the effective size of the pulse. Adding the factor to the isentropic formula and solving for the pump factor gives us the following equation:

For example, the pressure curve from an undampened, two-shoe, 2.5-inch peristaltic hose pump shows a sharp increase in flow, followed by a “no-flow” or negative flow zone. In this instance, the line has a ball valve that is creating the flow restriction for back pressure. The blue line shows the undampened pressure spikes (see Figure 2). The red line shows the pressure changes of the same pump with the same back pressure valve setting but now using a dampener. This sample dampener has an actual gas volume of 415 cubic inches, and the dampener is 90-percent gas filled. The base pressure is 14.15 psig, and the pulse is 76.9 cubic inches. If the pressure fluctuation is calculated using the isentropic pressure formula, the result is:

It is important to remember to add 14.7 psi to convert from gauge to absolute pressure, then subtract 14.7 psi again to get the final result in gauge pressure. This pump setup was tested, and the actual pressure variation was determined to be 7.38 psi. Therefore, the result is:

If the example above is used and it is decided that a pressure fluctuation of 15 psi would be acceptable, the formula with the previously calculated pump factor can be used to determine what size of dampener is needed.

Table 1 lists some approximate pump constant factors that can be used when sizing dampeners for different pump types. These factors are approximate, and the results may vary significantly with the many variables involved.

A triplex plunger pump doses methanol, which is metered on the discharge side. Without a dampener to control pulsations and smooth out the flow, the installed flow meters were giving inaccurate readings.

When using a triplex pump, all three chambers of the pump must stay full of fluid with no voids. Any voids or pockets can cause seal leakage, pump vibration and excess pump noise.

The solution was to install a pulsation dampener at the pump discharge to smooth the flow and remove pressure pulsations. This allowed the dosing to be more accurate. An inlet stabilizer (suction dampener) was also installed on the inlet side of the pump to act as an accumulator to keep the pump chambers filled. The inlet stabilizer also removed pulsations created by the pump on its inlet stroke. Both devices were sized based on the pump type, flow rate and operating pressure.

During the filling of a drum with a flexible hose, an automatic valve would close and cause a water hammer effect. All the pipes leading into the system would shake until they broke loose from their supports. The solution was to install a pulsation dampener at the beginning of the flexible hose connection.

The pulsation dampener was sized based on the flow parameters and installed at the beginning of the flexible hose. When the automatic valve closed, the hose and pulsation dampener effectively absorbed a portion of the water hammer, eliminating pipe shake and improving operational safety.

The sizing of a pulsation dampener is critical to achieving the desired result. Finding and using the correct constant pump factor in dampener sizing is a key part of the solution. As long as the pulsation dampener is properly sized, positioned and charged, it will effectively dampen pulsations to protect equipment and keep the pressure pulses within design parameters.

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.

The erosive and/or damaging effects of drilling fluid flow through a valve can damage the seating surfaces so that the ideal flat-surface-to-flat surface seating is not achieved. Also, the drilling fluid can damage a guide (e.g. ribs and a channel for receiving a stem or rod projecting from a valve member) or guide fingers so that the ideal surface seating is not achieved. In some instances, damage to a guide or to guide fingers results in a flat valve member surface contacting a flat seating surface at an angle so that effective valve closure is not possible or so that the valve is insufficiently closed for efficient operation. In some aspects, erosive drilling fluid flow renders initially-flat seating surfaces non-flat with resulting ineffective sealing and valve closure.

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 for pumping a drilling fluid mixture, the drilling fluid mixture containing drilling fluid and solids, the system having: a pump apparatus; the pumping apparatus having a body with 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 within the body for inhibiting pulsations of fluid pumped from the pump apparatus In certain valves according to the present invention a valve actuator is used which is pneumatically powered without certain mechanically moving parts used in prior valves.

Accordingly, the present invention includes features and advantages which are believed to enable it to advance pumping system technology. Characteristics and advantages of the present invention described above and additional features and benefits will be readily apparent to those skilled in the art upon consideration of the following description of preferred embodiments and referring to the accompanying drawings.

Certain embodiments of this invention are not limited to any particular individual feature disclosed here, but include combinations of them distinguished from the prior art in their structures, functions, and/or results achieved. Features of the invention have been broadly described so that the detailed descriptions of embodiments preferred at the time of filing for this patent that follow may be better understood, and in order that the contributions of this invention to the arts may be better appreciated. There are, of course, additional aspects of the invention described below and which may be included in the subject matter of the claims to this invention. Those skilled in the art who have the benefit of this invention, its teachings, and suggestions will appreciate that the conceptions of this disclosure may be used as a creative basis for designing other structures, methods and systems for carrying out and practicing the present invention. The claims of this invention are to be read to include any legally equivalent devices or methods which do not depart from the spirit and scope of the present invention.

What follows are some of, but not all, the objects of this invention. In addition to the specific objects stated below for at least certain embodiments of the invention, other objects and purposes will be readily apparent to one of skill in this art who has the benefit of this invention"s teachings and disclosures. It is, therefore, an object of at least certain preferred embodiments of the present invention to provide new, useful, unique, efficient, nonobvious dampener systems for drilling fluid pumping systems and methods of their use;

The present invention recognizes and addresses the problems and needs in this area and provides a solution to those problems and a satisfactory meeting of those needs in its various possible embodiments and equivalents thereof. To one of skill in this art who has the benefits of this invention"s realizations, teachings, disclosures, and suggestions, various purposes and advantages will be appreciated from the following description of certain preferred embodiments, given for the purpose of disclosure, when taken in conjunction with the accompanying drawings. The detail in these descriptions is not intended to thwart this patent"s object to claim this invention no matter how others may later attempt to disguise it by variations in form, changes, or additions of further improvements.

The Abstract that is part hereof is to enable the U.S. Patent and Trademark Office and the public generally, and scientists, engineers, researchers, and practitioners in the art who are not familiar with patent terms or legal terms of phraseology to determine quickly, from a cursory inspection or review. the nature and general area of the disclosure of this invention. The Abstract is neither intended to define the invention, which is done by the claims, nor is it intended to be limiting of the scope of the invention or of the claims in any way.

It will be understood that the various embodiments of the present invention may include one, some, or all of the disclosed, described, and/or enumerated improvements and/or technical advantages and/or elements in claims to this invention.

Certain aspects, certain embodiments, and certain preferable features of the invention are set out herein. Any combination of aspects or features shown in any aspect or embodiment can be used except where such aspects or features are mutually exclusive. BRIEF DESCRIPTION OF THE DRAWINGS

A more particular description of embodiments of the invention briefly summarized above may be had by references to the embodiments which are shown in the drawings which form a part of this specification. These drawings illustrate embodiments preferred at the time of filing for this patent and are not to be used to improperly limit the scope of the invention which may have other equally effective or legally equivalent embodiments.

FIG. 2F is a perspective view, partially cutaway, of a pump module according to the present invention with valve assemblies according to the present invention.

Certain embodiments of the invention are shown in the above-identified figures and described in detail below. Various aspects and features of embodiments of the invention are described below and some are set out in the dependent claims. Any combination of aspects and/or features described below or shown in the dependent claims can be used except where such aspects and/or features are mutually exclusive. It should be understood that the appended drawings and description herein are of certain embodiments and are not intended to limit the invention or the appended claims. On the contrary, the intention is to cover all modifications, equivalents and alternatives falling within the spirit and scope of the invention as defined by the appended claims. In showing and describing these embodiments, like or identical reference numerals are used to identify common or similar elements. The figures are not necessarily to scale and certain features and certain views of the figures may be shown exaggerated in scale or in schematic in the interest of clarity and conciseness.

As used herein and throughout all the various portions (and headings) of this patent, the terms “invention”, “present invention” and variations thereof mean one or more embodiments, and are not intended to mean the claimed invention of any particular appended claim(s) or all of the appended claims. Accordingly, the subject or topic of each such reference is not automatically or necessarily part of, or required by, any particular claim(s) merely because of such reference. So long as they are not mutually exclusive or contradictory any aspect or feature or combination of aspects or features of any embodiment disclosed herein may be used in any other embodiment disclosed herein. DETAILED DESCRIPTION OF THE INVENTION

The system 500 shown in FIG. 1 includes a derrick 502 from which extends a drillstring 504 into the earth 506. The drillstring 504, as is well known, can include drill pipes and drill collars. A drill bit 512 is at the end of the drillstring. A rotary system 514, top drive system 526, and/or a downhole motor 532 (“fluid motor”, “mud motor”) may be used to rotate the drillstring 504 and the drill bit 512. A typical drawworks 516 has a cable or rope apparatus 518 for supporting items in the derrick 502. A mud pump system 522 according to the present invention with one, two, three-to-ten, or more mud pumps 521 according to the present invention each with pumping modules with one or two valves according to the present invention supplies drilling fluid 524 to the drillstring 504. Drilling forms a wellbore 530 extending down into the earth 506. Each mud pump 521 has at least one valve 501 according to the present invention or (as shown in FIG. 1A schematically) multiple pumping modules 503 each with a suction valve 505 according to the present invention and a discharge valve 506 according to the present invention. Each mud pump 521 has a main crank shaft 521 c.

During drilling, the drilling fluid 524 is pumped by pump(s) 521 of the mud pump system 522 into the drillstring 504 (thereby operating a downhole motor 532 if such an optional motor is used). Drilling fluid 524 flows to the drill bit 512, and then flows into the wellbore 530 through passages in the drill bit 512. Circulation of the drilling fluid 524 transports earth and/or rock cuttings, debris, etc. from the bottom of the wellbore 530 to the surface through an annulus 527 between a well wall of the wellbore 530 and the drillstring 504. Cuttings and debris are removed from the drilling fluid 524 with equipment and apparatuses not shown, and it is re-circulated from a mud pit or container 528 by the pump(s) of the mud pump system 522 back to the drillstring 506. Also, some desirable solids may be added to the drilling fluid.

A system 10 according to the present invention as shown in FIGS. 2A and 2B has a main housing 12 mounted on a base 8 with an optional crane system 20 for lifting and moving system parts. A pedestal 21 of the crane system 20 is rotatably mounted on a bearing assembly 22 on the housing 12. A lift apparatus 23 is movably mounted on a beam 24 and a support 25 extends down from the lift apparatus 23. A chain hoist lift may be used with the structure shown which is attached to the support 25. Motors 14 each drive pinions 16 which in turn drive a drive gear 18 (see FIG. 3C) to move pistons 19 for six removable pump modules 650 (as described below; may be any module disclosed herein and/or may have any valve assembly or valve assemblies disclosed herein). A pressure relief apparatus (e.g. one or more relief valves) is provided for the modules 650 and, as shown, in one aspect, for each of the six modules 650 there is a pressure relief valve 13. Optional rails 15 project up from the housing 12.

An oil pump 2 pumps lubricating oil to various parts of the system. A water pump 4 pumps water to a filtration system (not shown) and a cooler (not shown). The pumps are mounted on pump mounts 8 bconnected to the base 8. Doors 3 and 5 (one each for each pump system 30) provide access to various internal parts of the system 10. Drilling fluid enters the system 10 through an inlet 7 and is pumped out via the modules 650 to a main outlet 9.

The modules 650 have a body 602 with a first bore 602 aand a second bore 602 b. A discharge valve assembly according to the present invention is in the first bore and a suction valve assembly according to the present invention is in the second bore. With a piston fluid is pumped into a chamber 652 of the module 650 via an inlet port 604 and is discharged from the module 650 into a discharge conduit 634 via an outlet port 606.

FIG. 2F shows the relative positions of two valve assemblies 100 a, 100 b(like the valve assembly 100) according to the present invention as they are present in a block of a mud pump module. The valve assemblies 100 a, 100 b(which may be any valve assemblies disclosed herein) are in bores 642, 643, respectively, in a block 644. The block 644 can be used in a system like that of FIG. 2A.

FIGS. 2G-2I show two valve assemblies 100 x, 100 y(like the valve assembly 100 a, FIG. 9A; may be any valve assembly according to the present invention) as they are disposed in a block B (shown in dotted line; may be any suitable block or body; including, but not limited to, the body 602 or block 644 referred to above) of a mud pump system. Fluid is sucked in by action of the suction valve assemblies 100 xthrough a suction inlet 400 and discharged by action of the discharge valve assembly 100 ythrough a discharge outlet 402. The fluid is received in a pumping chamber 404.

Fluid pumped from the chamber 404 can impact parts of the discharge valve 100 x. Optionally, an accumulator/dampener 410, positioned within the block B, is in fluid communication with the pumping chamber 404. The accumulator/dampener 410 reduces undesirable pulsations of fluid under pressure from the pumping chamber 404. Any suitable known accumulator/dampener may be used.

FIGS. 3A and 3B show a valve assembly 100 according to the present invention which can serve as a suction valve or a discharge valve for a mud pump system (e.g., but not limited to, the suction valve assembly 680 and the discharge valve assembly 630 described above; or the suction valve 100 xand the discharge valve 100 ydescribed above). FIG. 4 shows top portions of the valve assembly 100.

The valve assembly 100 has a hollow cartridge stem 102 with an interior channel 104 within which are located a valve actuator 130 and an adapter 106. A spring support 108, connected to a flange 110 of the cartridge stem 102, has an end 112 which is encompassed by part of an expansion spring 120 an end of which abuts the spring support 108.

A poppet (or curved valve member) 114 rests on a support 116. An end 122 of the spring 120 abuts and is biased against a bottom of the support 116. A ball 118 rests on a ball support 124 which rests on the support 116. A cable 128 (i.e. a non-rigid connector) (made of any known cable material) connected to the ball 118 passes through a hole 140 in and through the support 124, through a hole 142 in the support 116, through the spring 120, through a hole 143 in the spring support 108, through a hole 144 in the adapter 106 which is and is connected to the adapter 106 connected to an actuator 130.

A washer 151 above the ball 118 abuts an underside 115 of the poppet 114. A recess 152 within the poppet 114 houses the ball 118, the washer 151 and the support 124. The poppet 114 has a tapered surface 136 for sealingly abutting a valve seat and a seal of a valve seat as described below.

The poppet 114 is movable toward and away from a valve seat 160. The valve seat 160 has a channel 162 for fluid flow therethrough. The poppet 114 selectively closes off and opens up the channel 162 to fluid flow. Part of the channel 162 is sized and configured for the poppet 114. A surface 166 of the valve seat 160 is positioned to seal against the tapered of the surface 136 of the poppet 114. Optionally, there are no guide fingers projecting from the poppet 114 (although it is within the scope of the present invention to use them); and there are no arms or ribs across the valve seat (it is unobstructed) for receiving and stabilizing a rod, stem or neck projecting from a poppet; and there is no rod, neck or stem projecting from the poppet. Thus, flow through the channel 162 is unobstructed by such parts which are present in many prior valves.

A recess 168 around the valve seat 160 holds a seal 169. Part of the surface 136 of the poppet 114 sealingly abuts the seal 169 when the valve assembly is closed, preventing fluid flow. Thus dual sealing is achieved.

The poppet 114 has a range of freedom of movement within the channel 162 of the valve seat 160. However the poppet 114 is located within and with respect to the valve seat 160, part of the outer tapered surface 136 of the poppet 114 will sealingly abut the seal 169 and the surface 136 will sealingly abut the surface 166. The poppet 114 can be aligned (or not) with the valve seat 160, but either way an effective seal is maintained with part of the surface 136 sealed against the seal 169. Movement of the poppet 114 on the ball 118 and the sizing and configuration of the various parts contribute to permissible freedom of movement of the poppet 114 without sacrificing the sealing necessary to close the valve assembly.

FIG. 5 shows the valve actuator 130 which can be, in certain aspects, any suitable known controllable, valve actuator, e.g., but not limited to “muscle” apparatuses, pneumatic cylinder actuators, hydraulic cylinder actuators, and electromagnetic actuators.

In one aspect, as shown in FIG. 5, the valve actuator 130 is a controlled, pneumatically powered actuator known as a FESTO (TRADEMARK) “muscle” actuator. The actuator 130 has an expandable hose 132 mounted between two bases 134, 135. Air under pressure is introducible into the interior of the hose 132 through a channel 137 in a pneumatic coupling 139. The upper base 134 is connected to an adapter support 127 to which the adapter 106 is secured.

As shown in FIG. 5, air under pressure has not yet been applied within the hose 132. Once air is applied the hose moves outwardly, effectively moving the top base 134 toward the lower base 135 and thereby pulling the adapter 106 to pull the cable 128 and move the poppet 114 out of sealing contact with the valve seat 160 against the force of the spring 120.

FIG. 6 shows one embodiment, a spring 120 a, of a spring 120. As compared to prior known spring designs, the spring 120 ahas a spring body with a smaller spring diameter, a, and with a higher spring force; but the wire diameter is relatively large, e.g. 0.22 inches, which results in the higher spring force. Use of an actuator like the actuator 130, FIG. 5, makes it possible to use a spring with the increased spring force (with the increased wire diameter). The overall diameter, b, of the spring 120 ais relatively smaller than prior springs because the spring 120 adoes not have to accommodate the relatively large necks of certain prior valve members. Certain prior mud pump valve springs reached a known resonant frequency (e.g. about 40 Hz to 43 Hz) creating poppet oscillations that resulted in an improperly seated poppet and in fluid pulsations transmitted downstream of a valve assembly. Due to its size and weight, the spring 120 ahas a higher natural frequency than those prior springs which resonate around 40 Hz and, thus, more force is required to resonate the spring 120 a. In certain aspects the spring 120 (or 120 a; or the spring 120 b, FIG. 7A) is sized and configured so its natural resonant frequency is about 25% higher than that of certain known springs (e.g., in one aspect 50 Hz vs 43 Hz). This reduces the chance of flow-induced resonance in the valve assembly with such a spring; provides better, more stable control of the valve assembly"s poppet; and provides more positive seating of the poppet against the valve seat.

FIGS. 7A and 7B show a spring 120 baccording to the present invention which has a spring body 120 cand an end tapered portion 120 dwhich abuts a support (e.g. like the support 116, FIG. 3A). The tapered portion 120 d, since it is narrower than a base 120 eof the spring 120 b, contributes to the freedom of movement of the poppet 114 (e.g. as in FIG. 8A).

FIGS. 8A and 8B illustrate steps in the operation of a valve assembly 100 (which has a spring 120 b, although any suitable spring may be used). As shown in FIG. 8A, air under pressure has not yet been applied within the hose 132 and the and the spring 120 burges the poppet 114 into sealing contact with the seal 169 and with the valve seat 160. The valve assembly 100 is closed to fluid flow therethrough. Fluid pressure also forces the poppet against the valve seat. On the discharge side of the valve seat at the beginning of the pumping/compression part of a cycle, the spring 120 band the fluid within a discharge manifold pushes the poppet 114 against the seat. This continues until the pressure within the discharge manifold drops below the pressure within the pumping cylinder and/or until the actuator 130 is commanded to open. On the suction side, the fluid within the pumping cylinder pushes the poppet 114 against the seat 160 again during the compression part and until the actuator 130 is commended to open the valve. When the “muscle” of the actuator 130 is not expanded, there is residual air trapped between the commanding valve and the actuator 130. The pressure of this trapped air is close to the pressure that existed in this line at the moment of exhausting the air and closing off the valve"s exhaust port. When the actuator is flexed, there is air at a pressure that is sufficient to open the valve, e.g. 110 psi. The actuator and air lines are filled in order to decrease the actuator"s response time—the time to respond to a commanding pressure. If the actuator is completely empty or, with, e.g. air at atmospheric pressure, it will take slightly longer for the actuator to respond, because when such a high pressure is applied the cavity would have to be filled with air first, then compress the air just introduced to a high enough pressure to barely stretch the hose 132 and only after that will the hose 132 change its length or respond to a commanding pressure.

As shown in FIG. 8B, air under pressure from an air supply 200 (with a proportional control valve 200 p) has been applied within the hose 132 causing it to expand and pulling the cable 128 away from the valve seat 160. In so doing, the poppet 114 is moved out of sealing contact with the seat 160 and the seal 169 of the valve seat 160 and the valve assembly is opened to fluid flow permitting fluid to flow into and out from a mud pump module housing the valve assembly.

It is advantageous that the poppet is part of the valve cartridge. During assembly, when the pump is assembled for the first time, it is much easier to have a preassembled valve cartridge and, without adjustments, to insert and bolt it in and have it immediately become functional. Moreover, in servicing the valve, it is much easier to extract the entire cartridge, versus bits, individual parts, and/or pieces. In certain current designs, a poppet/valve has a pseudo cartridge design in the sense that the valve has no restricting elements to keep it attached to the cartridge. In other words, the cartridge can be loosely put together prior to assembly and it can be inserted as a cartridge being secured to the body by bolts. However, if during this assembly process, or later on during servicing the valve, this cartridge is turned upside down, the valve itself can become loose and fall to the ground.

Often in such prior systems there is no element like a snap ring to secure the valve to the cartridge. It is also advantageous that the seal is part of the valve housing. It is easier to have the seat part of a block that can be preassembled to the pump and, later on, during a later step in manufacturing, to bolt on to it a subassembly like the valve cartridge.

In designs according to the present invention, seals, e.g. the seal 169, do not resonate. According to the present invention, such seals are surrounded by a support and have no extraneous or “banging” features which could be excited by a surrounding flow stream.

In certain aspects according to the present invention, poppets and seats are made of ceramics which do not rust. In certain particular aspects, an alumina based ceramic offers very high strength and good wear resistance. In other aspects, a boron carbide ceramic can be used which has excellent erosion wear resistance. Both of these two ceramics have a higher erosion resistance then steel. In certain aspects the poppets of assemblies according to the present invention are made with a steel core surrounded by a ceramic. The steel core supports the Belleville washers and can have cut threads into it. A ceramic outer skin provides erosion resistance. In certain aspects, the special profiles facilitate the flow opening and closing the valve gradually.

In certain current designs, valves have two parallel surfaces. Often these surfaces form a seal that is part of conical bodies; i.e. the seal has a conical machined surface against which is pushed a poppet. The poppet"s sealing surface is also conical so that, at every instance, the seat"s and poppet"s sealing surfaces are parallel. During discharge, when the two bodies are separating and, thus, allowing the fluid to flow from the pumping chamber into the discharge manifold, the fluid is squeezed in between these flat surfaces. During this phase the fluid"s velocity can be greatly increased as it passes from a large cross section of the pumping chamber into a small one with parallel surfaces of the valve"s passage way. Moreover, because there is no controlling actuator, such a valve can open suddenly when the fluid"s pressure exerts onto the valve"s face a force slightly higher than that developed by the spring acting on the opposite face. As the fluid leaves at high velocity, it enters into a larger cross section that is the discharge manifold The high velocity and energy fluid acts almost like a piston in this case and pushes an adjacent block of fluid along the discharge line. This sudden move of a significant block of fluid can create a “bang” or a specifically loud noise almost like a pounding. This repeated banging/pounding can have detrimental effects on the drill line or other equipment.

In certain valve assemblies according to the present invention, the flat parallel surfaces are replaced by curved ones. Additionally, there is a controlling actuator that can open the valve before pressure in the pumping chamber reaches a value high enough to counteract the spring and, thus, to open the vale. Pressure at which the fluid leaves the pumping chamber is greatly reduced. Being formed in between two curved surfaces, the valve"s passage way flow characteristics do not impart a high velocity/energy to the fluid stream. Consequently, the fluid enters and leaves the discharge manifold and line respectively in a more dispersed manner. There is no “bang” as in certain previous valves because the fluid does not flow in discrete “blocks”.

The control system CS controls the air supply 200 and, thus, controls the valve assembly 100. This is in contrast to prior valves in which fluid flow opens and closes the valve. In one aspect, the control system controls the speed with which the parts move and thereby controls the speed of opening and of closing off the valve. Using appropriate software programming of programmable media in the control system, the control system controls an electro proportional valve control (e.g. the valve 200 p. FIG. 8B) that, in turn, controls the amount of air that enters or leaves the actuator 132. Consequently, the control system controls how fast, how long and how much the valve is opened. Gradual opening and closing is possible which reduces pressure pulsations. Each pump shaft (crankshaft) may have a speed sensor in communication with the control system (e.g. a sensor 521 s, FIG. 1). In systems with electric motors that drive the crankshaft(s), the motors are commanded through software in the control system and the same speed control signal can be broadcast to the control system. A dedicated speed sensor or a linear displacement transducer installed in every cylinder provides information for a closed loop control system (usable, e.g., to diagnose a pump in case of failure). With valve assemblies according to the present invention, the valves are not connected to the crankshaft.

The control system has programmable media, e.g. in a computer, computers, and/or PLC(s). In one aspect, the control system is preloaded with a program that includes a defining equation and a curve fitter. The defining equation is a function of pump shaft speed. The curve fitter compares the curve generated by the defining equation with an “ideal” curve desired to drive the valve The ideal curve usually represents the valve"s speed, or acceleration, or opening and/or, a different relevant parameter plotted versus time. The output from the control system drives a proportional valve, a valve that controls the actuator 130, e.g., in one aspect, supply air into a FESTO (TRADEMARK) “muscle”. Thus, the valve being actuated closely follows the preprogrammed curve/equation and the valve opens or closes at a certain velocity or acceleration, or that it opens at a certain rate over the duration of a pumping cycle. The opening or closing rate can be constant or variable. That is, the valve can start opening at a certain low rate followed by a higher rate followed by a different rate, and so on.

In one aspect, during a cycle the valve tends to follow a certain bell-shaped curve. Thus, the valve starts opening at a low rate followed at the very next instance by a slightly higher rate and in the next instance by an even higher rate and so on. All this is followed on the descending side of the curve by a lower rate followed by a slightly lower rate and so on until the valve closes. By introducing or expelling fluid into or from the pumping chamber at certain times the pump"s behavior is changed or the pump"s flow is measurable.

The mechanical equivalent of controlling a valve"s opening rate is a cam. The cam, through its profile, controls how fast and in what relationship relative to another element, e.g. a crankshaft, the valve will open or close. In other words, it controls the valve"s rate (displacement versus time). However, a cam"s profile can not be changed very easily because it is cut in metal. A practical method is to introduce a hydraulically actuated push rod or cam follower in between the cam and valve. Thus, the rate can change at will within a limited range. In the control strategy according to the present invention there is no piece of hardware/cam that limits the valve"s rate. Consequently, in the proposed actuation and control strategy, the desired curve can be changed on the fly as long as the controller, e.g. a computer or PLC, can accept/support it. Programmability makes this equivalent to an infinitely variable profile cam shaft and the pump"s output flow and vibration can be controlled. (An undesirable consequence of output flow in certain prior systems is component failure, e.g. due to cavitation.)

With the curved mating sealing surfaces of the valve seat and poppet, any contact results in an effective seal. Pressure fluctuations generated in or by prior art valves are reduced or eliminated and valve control reduces pressure fluctuation in the discharge line during pump operation.

Systems according to the present invention provide a fail safe mode. If a valve assembly according to the present invention that is inserted fails, then, for safety reasons, the pump continues working at either reduced or normal parameters until it is safe to stop it for service. In systems according to the present invention, if the actuator fails, e.g. if the muscle fails, it breaks or bursts, the valve will operate unrestricted (e.g. as a current known design valve). Thus, the pump can continue working at almost the same parameters until it is safe to stop it.

FIGS. 9A and 9B show a valve assembly 100 a, like the valve assembly 100 (like numerals indicate like parts) with a spring 120 band a poppet 114 a. The poppet 114 ahas a nose 114 nprojecting from a poppet body 114 b. The nose 114 nprojects into the flow channel 162 of the valve seat 160. In certain aspects, in systems according to the present invention the surface on the valve seat becomes, advantageously, more elastic. In a seal, two surfaces or edges are pushed against each other by a force. This acting force can be perpendicular to or at an arbitrary angle relative to the sealing surfaces. In systems according to the present invention the sealing bodies are the rubber seal and the poppet in one instance and, the seat itself and the poppet in a second instance. During a valve closing cycle, the first seal occurs in between a rubber O-ring and poppet. The acting force is axial relative to the poppet, but it is at an angle relative to the edge of contact between the two curved surfaces of the O-ring and poppet respectively. When the two bodies come into contact, at the point of contact, the vector components of this acting force are a normal to curved surfaces component and a tangential to curve components. This tangential component will stretch the rubber (the over hanging part of it) instead of purely compressing it. With the rubber O-ring being surrounded/supported by the seat"s rigid body, the rubber will take a very high force in compression as the normal-to-curved surfaces vector component. The rubber becomes difficult to compress when it is surrounded by a rigid wall. Thus a mechanical maze is formed and, thus, the fluid encounters a high flow resistance. There is a sequence of high pressure (inside the pumping chamber), followed by a no flow area (where the rubber O-ring contacts the poppet), followed by a low pressure area (right after the rubber seal) and finally, followed by a no flow area at a contact between the poppet and the seat. Also, the shape of the deformed rubber O-ring at the leading edge toward the impinging fluid does not allow the fluid to enter in between the poppet and seal.

Valve “shivering” occurs when a valve is not actuated (pushed or pulled onto its seat) with a high enough force, and flow induced forces fully or partially unseat or seat the valve in a rapid sequence. Thus, the valve can not fulfill its primary function of separating two c