Content of PetroWiki is intended for personal use only and to supplement, not replace, engineering judgment. SPE disclaims any and all liability for your use of such content. More information
Difference between revisions of "PEH:Sucker-Rod Lift"
|Line 17:||Line 17:|
== Introduction ==
== Introduction ==
<br/>This chapter discusses the specific artificial-lift technique known as beam pumping, or the sucker-rod-lift method. Many books, technical articles, and industry standards have been published on the sucker-rod lift method and related technology.<ref name="r1">_</ref><ref name="r2">_</ref><ref name="r3">_</ref><ref name="r4">_</ref><ref name="r5">_</ref><ref name="r6">_</ref><ref name="r7">_</ref> This chapter is a complete revision of previous editions of the ''Petroleum Engineering Handbook'',<ref name="r6">_</ref> but it combines the prior three relevant chapters that covered downhole rod pumps and sucker rods, along with pumping units and prime movers. Additionally, the other components of a sucker-rod pumping installation are discussed, including applicable engineering and operating information. The complete operating system should be understood and addressed to properly design, install, and operate this or any other type of artificial-lift system. Thus, this chapter uses the Gipson and Swaim "Beam Pump Design Chain" as a foundation and builds on this design philosophy by using relevant, published technology and the latest industry practices.<ref name="r5">_</ref><ref name="r6">_</ref><ref name="r7">_</ref>
<br/>This chapter discusses the specific artificial-lift technique known as beam pumping, or the sucker-rod-lift method. Many books, technical articles, and industry standards have been published on the sucker-rod lift method and related technology.<ref name="r1">_</ref><ref name="r2">_</ref><ref name="r3">_</ref><ref name="r4">_</ref><ref name="r5">_</ref><ref name="r6">_</ref><ref name="r7">_</ref> This chapter is a complete revision of previous editions of the ''Petroleum Engineering Handbook'',<ref name="r6">_</ref> but it combines the prior three relevant chapters that covered downhole rod pumps and sucker rods, along with pumping units and prime movers. Additionally, the other components of a sucker-rod pumping installation are discussed, including applicable engineering and operating information. The complete operating system should be understood and addressed to properly design, install, and operate this or any other type of artificial-lift system. Thus, this chapter uses the Gipson and Swaim "Beam Pump Design Chain" as a foundation and builds on this design philosophy by using relevant, published technology and the latest industry practices.<ref name="r5">_</ref><ref name="r6">_</ref><ref name="r7">_</ref>
=== Beam-Pumping Systems ===
=== Beam-Pumping Systems ===
Revision as of 14:01, 6 October 2020
Petroleum Engineering Handbook
Larry W. Lake, Editor-in-Chief
Volume IV - Production Operations Engineering
Joe Dunn Clegg, Editor
Copyright 2006, Society of Petroleum Engineers
Chapter 11 – Sucker-Rod Lift
This chapter discusses the specific artificial-lift technique known as beam pumping, or the sucker-rod-lift method. Many books, technical articles, and industry standards have been published on the sucker-rod lift method and related technology. This chapter is a complete revision of previous editions of the Petroleum Engineering Handbook, but it combines the prior three relevant chapters that covered downhole rod pumps and sucker rods, along with pumping units and prime movers. Additionally, the other components of a sucker-rod pumping installation are discussed, including applicable engineering and operating information.. The complete operating system should be understood and addressed to properly design, install, and operate this or any other type of artificial-lift system. Thus, this chapter uses the Gipson and Swaim "Beam Pump Design Chain" as a foundation and builds on this design philosophy by using relevant, published technology and the latest industry practices.
Beam-Pumping SystemsBeam pumping, or the sucker-rod lift method, is the oldest and most widely used type of artificial lift for most wells. A sucker-rod pumping system is made up of several components, some of which operate aboveground and other parts of which operate underground, down in the well. The surface-pumping unit, which drives the underground pump, consists of a prime mover (usually an electric motor) and, normally, a beam fixed to a pivotal post. The post is called a Sampson post, and the beam is normally called a walking beam. Fig. 11.1 presents a detailed schematic of a typical beam-pump installation.
This system allows the beam to rock back and forth, moving the downhole components up and down in the process. The entire surface system is run by a prime mover, V-belt drives, and a gearbox with a crank mechanism on it. When this type of system is used, it is usually called a beam-pump installation. However, other types of surface-pumping units can be used, including hydraulically actuated units (with and without some type of counterbalancing system), or even tall-tower systems that use a chain or belt to allow long strokes and slow pumping speeds. The more-generic name of sucker-rod lift, or sucker-rod pumping, should be used to refer to all types of reciprocating rod-lift methods.
Linked rods attached to an underground pump are connected to the surface unit. The linked rods are normally called sucker rods and are usually long steel rods, from 5/8 to more than 1 or 1¼ in. in diameter. The steel rods are normally screwed together in 25- or 30-ft lengths; however, rods could be welded into one piece that would become a continuous length from the surface to the downhole pump. The steel sucker rods typically fit inside the tubing and are stroked up and down by the surface-pumping unit. This activates the downhole, positive-displacement pump at the bottom of the well. Each time the rods and pumps are stroked, a volume of produced fluid is lifted through the sucker-rod tubing annulus and discharged at the surface.
Selecting the Sucker-Rod Pumping MethodMany factors must be considered when determining the most appropriate lift system for a particular well. The chapter on Artificial Lift Selection in this volume of this Handbook presents a discussion of the normally available artificial-lift techniques, their advantages and disadvantages, and the selection of a method for a well installation.
Because of its long history of successfully lifting well fluids, the sucker-rod lift method is normally considered the first choice for most onshore, and even some offshore, installations all over the world. This method is limited by the size of the casing, tubing, and downhole pump; the strength and size of the various rods; and the speed with which they can be reciprocated. Under favorable conditions, approximately 150 BFPD can be lifted from greater than 14,000 ft, while more than 3,000 BFPD can be lifted from less than 2,000 ft. Some of the major advantages and disadvantages of this lift technique are shown in Table 11.1.
The Producing Reservoir
Understanding the makeup of the producing reservoir, its pressure, and the changes that occur in it are important to attain maximum production. Because reservoir conditions change as fluids are produced, ongoing measurement of the reservoir conditions is necessary. The main considerations in measuring and understanding the reservoir are the types and volumes of reservoir fluids being produced, their pressures in both the reservoir and at the wellbore or pump intake, and the effects these fluids have as they pass through the producing system.
The relationship between the reservoir-fluid inflow and the produced-fluid outflow is extremely important for any artificial-lift method. This should be monitored and controlled so that any excessive damage to the lift equipment is avoided while profitably obtaining the maximum amount of fluids. Undesirable effects result when the producing equipment's capacity is not properly balanced with reservoir-fluid inflow. These effects include the following:
- Loss or deferment of production.
- Excessive producing costs.
- Premature equipment failure.
- Ineffective use of energy.
- Increased operating expenses.
A variety of well tests and measurements may be used to determine production rates for oil-, gas-, and water-supply wells and to observe the status of the reservoir. Each test reveals certain information about the well and the reservoir being tested. The main reservoir considerations are determining bottomhole pressure and the inflow relationship of the fluids with changing reservoir and pump-intake pressure.
Bottomhole-pressure-measuring equipment (pressure bombs) makes it possible to determine reservoir and tubing intake pressures within the desired range of accuracy. When this test is conducted at scheduled intervals, valuable information about the decline or depletion of the reservoir from which the well is producing can be obtained. However, it is difficult to obtain either bottomhole reservoir or operating pressures while the rod-pump system is installed and operating.
Calculations of the bottomhole pressures can be obtained by using instruments that detect the fluid level in the casing/tubing annulus. The simplest instrument is a fluid-level sounder with a strip chart. Bottomhole pressures can be estimated from the gravity of the fluids (i.e., oil, water, and gas), the volumes produced, and the fluid level. If producing and shut-in conditions are known, then approximate producing and shut-in reservoir pressures can be determined.
The key to accurate bottomhole-pressure determination in any pumping well is the ability to predict the gradient of the fluid in the casing/tubing annulus. In 1955, W.E. Gilbert* developed an iterative calculation procedure on the effect of gas bubbling up a static fluid column. This can be used in a trial-and-error method to determine a gradient correction factor (F) to determine the pressure at the desired depth in the presence of gas production. If the term Q/(aP)0.4 is greater than 0.25, this method should be used with caution because this is an indication that liquid flow up the annulus may occur. Also, the crude pressure/volume/temperature (PVT) characteristics alter the results. The Gilbert curve and a calculation example are presented in "The Beam Pump Design Chain." 
Currently, the same fluid-level sounder equipment can be interfaced with a computer to determine the downhole pressures more easily. However, there still needs to be verification of the fluid-level indication to ensure that "false" or incorrect annulus fluid levels are not recorded. Additionally, the fluid gravities and produced volumes must be accurate and reflect actual conditions.
Knowing the reservoir and pump-intake pressures during static and operating conditions will allow a determination of the well's production capacity. This is required to optimize the artificial-lift equipment and properly size the equipment that is installed. The well productivity under varying production conditions must then be known.
Inflow Performance Relationship (IPR)
One of the most critical decisions in an artificial-lift system is the selection and design of equipment appropriate for the volume of fluid the reservoir produces. Other chapters of this Handbook detail the productivity index and IPR of fluids with changes in reservoir pressure. Because most fluid produced by an artificial-lift method is not single phase, it is not in a steady-state condition. Also, because most pumping operations occur after the fluid is below the bubblepoint pressure, the IPR method is usually considered. This technique takes into account various fluid phases and flow rates. It was originally devised by Vogel and described by Eickmeier. Each revision increased the accuracy of estimating flow rates from a well.
In the design of an artificial-lift system, it is necessary not only to predict production of the various fluids during existing conditions and reservoir pressure, but also to make a second type of prediction: future pressure performance. This can be accomplished with the IPR method and multiple, or a family of, IPR curves. Furthermore, the family of curves can be used to predict estimates of fluid production increases if the reservoir is repressurized from waterflooding or other secondary or tertiary methods.
Producing rates can be estimated within the desired range of accuracy using the IPR technique with two stabilized producing rates and corresponding stabilized producing pressures. This makes it possible to use the IPR without needing to shut in the well and lose production to obtain shut-in information. Obtaining a bottomhole pressure equal to 10% of the shut-in reservoir pressure is recommended for determining maximum production rates for sucker-rod lifted wells. At this pressure, the maximum well productivity will be 97% of the well's theoretical maximum production rate. However, the maximum lift-design rate should, in most cases, be slightly higher to permit some downtime and decreased pump efficiency.
In any artificial-lift system, the volume of gas produced should be considered in designing the system and in analyzing the operation after the system has been installed. A complete analysis requires knowing the volume of gas in solution, the volume of free gas, the formation volume factors, and whether gas is produced through the pump or is vented. If PVT analyses of reservoir fluids are available, they are the most accurate and easiest to use as a source of solution gas/oil ratio (GOR), formation volume factors, etc. The next best source is an analysis from a nearby similar reservoir.
A means of estimating PVT data is contained in Volumetric and Phase Behavior of Oil Field Hydrocarbon Systems. With the produced GOR, gas gravity, oil gravity, and reservoir temperature, the following can be estimated using the instructions included on each chart:
- Chart 1: The formation volume factor for the gas plus the liquid phases.
- Chart 2: The bubblepoint pressure.
- Chart 3: The formation volume factor of the bubblepoint liquid.
Gas VentingWhen pumping through tubing in the absence of a production packer, free gas, which breaks out of the oil, should be vented up from the casing/tubing annulus. However, when it is necessary to produce from beneath a production packer, a vent string can be installed. The possibility of needing a vent string should be considered when planning casing sizes for a new well.
Both the size of the vent string and the location of its bottom, with respect to the location of the pump intake and producing perforations, will influence the string's effectiveness in removing free gas. The string's diameter should be designed to allow the production of the anticipated free-gas volume with a pressure drop no greater than the desired producing bottomhole pressure minus the surface backpressure. If the required pressure drop is greater than this, a portion of the free gas will have to go through the pump. Fig. 11.2 is an indication of the effect of vent-string size on the pressure drop through it. Care should be taken if small-diameter tubing is used, because it may not allow all the gas to flow up the vent or may simply load up and prevent most gas flow.
Effects of Gas on Pump Performance
Gas that remains in solution when the liquid enters the pump increases the volume of total fluid through the pump compared to the liquid measured at the surface by the formation volume factor at pump-intake conditions. The gas also decreases the density of the fluid and, thus, the head or pressure to be pumped against in the tubing. Free gas that enters the pump must be compressed to a pressure equivalent to the head required to lift the fluid. This free gas will reduce the volume of both the produced liquid that enters the pump and the liquid measured at the surface. Any time the pump does not compress the free gas to a pressure greater than that exerted on the pump by the fluid column in the producing string, production ceases and the pump is said to be "gas locked." This condition can exist in both plunger and centrifugal pumps.
Intake PressureIntake pressure is the pressure in the annulus opposite the point at which the fluid enters the pump. If the pump-intake pressure is increased by increasing the pump submergence, the free-gas volume decreases because the fluid retains more gas in solution. Reducing the pressure drop in the pump-suction piping also reduces the free gas to be produced. The pump intake should not be deeper than is necessary to maintain the desired intake pressure. A pump intake that is too deep results in unnecessary investment and increased operating costs.
Fig. 11.3 is a graph of the liquid produced as a percent of the displacement of a plunger pump plotted against the pump-intake pressure for a typical reservoir. If the pressure is greater than the bubblepoint (Point A to B), the volumetric efficiency remains nearly constant. If all the gas can be vented rather than passed through the pump, the volumetric efficiency will increase as the formation volume factor decreases (Point B to C). If all the gas must be pumped, the volumetric efficiency decreases as the intake pressure drops to less than the bubblepoint (Point B to F). The lines B–D and B–E indicate the volumetric efficiency with a partial venting of gas as its presence declines. Note that the efficiency declines to a minimum at less than the bubblepoint and with further pressure reduction, starts to increase. A general conclusion is that to obtain better efficiencies, the pump-intake pressure should be maintained at or greater than the bubblepoint, or decreased to as low as possible to take advantage of the increased separation efficiencies at the low-pressure end. However, this considers only pump efficiency and not maximum production rate.
Gas bubbles entrained in the produced liquid(s) tend to rise because of the difference in the liquid and gas densities. The rate of bubble rise depends on the size of the bubbles and the physical properties of the fluid. The size of the bubbles increases as the pressure decreases. At low pump-intake pressures, the rate of gas-bubble rise in low-viscosity fluids will approximate 0.5 ft/sec, assuming a 400-μm bubble rise in water. The increase in bubble size and rate of rise as the pressure decreases causes the reversal in curves B–D and B–E in Fig. 11.3.
Downhole Gas Separators and AnchorsDownhole gas separators are used in gassy wells to increase the volume of free gas removed from the liquids before reaching the pump. However, they are not 100% effective in separating the gas. In sucker-rod-pumped wells, these separators are normally called "gas anchors." Gas anchors are usually designed and built in the field; Fig. 11.4 contains schematic drawings of six common types. The most commonly used are the "natural" gas anchor (A) and the "poor boy" gas anchor (C). Typically, there are two major components for these gas-anchor assemblies, the mud anchor run on the bottom of the tubing string and the dip tube or strainer nipple run on the bottom of the pump.
The largest downhole gravity separator is normally the casing/tubing annulus. This area provides a maximum down passage for liquid and up-flow area for gas. This allows the oil (and water) to move relatively slowly, typically, downward from the perforations to the pump, and permits the gas to separate and flow upward. For this reason, a natural gas anchor should be used whenever practical because it takes advantage of the entire casing internal cross-sectional area. This type of separator typically should be placed approximately 15 ft below the lowest most-active well perforations. However, if there is insufficient distance in the well to place the pump intake below the perforations, then the pump intake should be placed approximately 15 ft above the top-most perforation and a poor boy separator should be properly designed and installed.
There are limitations on how much gas can be handled by the downhole separator. If more gas is produced than can be handled by the separator, the gas will not separate completely. The downhole pump must then handle the excess gas. If the wells exceed these theoretical gas rates, then pump volumetric efficiency decreases, liquid production decreases, energy is wasted, and operating costs rise. The situation worsens if excessive gas enters the pump and there is insufficient compression ratio to pump all the fluids, resulting in a gas-locked pump. When this occurs, operating costs for this well increase dramatically because when there is no production, there is no revenue. However, a properly designed and spaced pump should not gas lock if the well is not pumped off.
Example calculations of the gas capacity of various casing/tubing annuli vs. different intake pressures have been presented in Hein. This reference also discusses the types of downhole separators and emphasizes the need to run a natural gas-anchor assembly whenever possible. Detailed discussions on design of the different types of separators, the arrangement of components, and example calculations for sizing components are presented by Gipson and Swaim. Improved gas separators with decentralized intakes have been introduced. This design aids in separation efficiency because it increases the local distance from the casing's inner diameter (ID) to the mud anchor, which results in an increased separation area. However, as with all specialty devices, the need to run this new design should be demonstrated by ensuring that the appropriate, standard systems have been properly installed and operated.
It is often recommended that the outside diameter (OD) of the gas anchors' steel mud anchor be less than the ID of the largest overshot or wash pipe that can be run in the well casing. This limits the gas-anchor separation capacity that can be secured in wells with small casings. Reinforced plastic mud anchors that can be drilled up, or steel designs that can be recovered with spears, should be considered when mud anchor OD must approach casing-drift diameter. This design would then be considered the "modified poor boy." Agreement should be obtained from the field before installation to ensure acceptance of the possible problems when trying to pull this type of installation.
Downhole Sucker-Rod Pumps
There are seven major components for downhole rod pumps: standing and traveling valves, plunger, barrel, seating assembly, pull tube or valve rod (for insert pump), and the fittings that hold the assembled pump together. The most common of these components and the final types of assembled pumps are covered by American Petroleum Inst. (API) Specification 11AX.
Types of Pumps
API recognizes two main types of pumps: rod and tubing. Rod pumps also are called insert pumps because they are run (inserted) in the production tubing. Tubing pumps are so named because the working barrel of this pump is coupled with the production-tubing string.
There is a wide range of plunger (or pump-bore) sizes standardized by the industry. The API pump-bore sizes that are currently available range from 1∕161 to 3¾ in. in diameter. This 1∕161 -in. size has been added back in the latest edition of the standard. Additionally, a new barrel type has been accepted in the latest API Spec. 11AX. This is the "X-type" barrel. It has a thin-walled barrel configuration for threads on either end of the heavy-walled barrel and is available for metal plungers only. This type of pump does not require the extension couplings normally needed for heavy-walled barrel pumps. Thus, this pump reduces the burst or collapse concerns of the thin-walled extension couplings and allows deeper producing depths to be attained.
API Pumps and NomenclatureWhile there are only two main types of pumps standardized by API, there are four different types of rod pumps. These are classified by the type of barrel (standing or traveling) and where the pump is anchored (top or bottom). Table 11.2 shows the letter designations for the various types of rod and tubing pumps that are available for different barrel thicknesses and either metal or soft-packed plungers.
The complete pump designation of an API pump adds dimensional diameters and lengths to the letter designations. This has been modified in the latest revision to incorporate all approved sizes and barrel types along with separating the extensions into the top and bottom lengths, if required. The complete API designation includes the following:
- Nominal tubing size (from 1.9- to 4.5-in. OD).
- Basic bore diameter (from 1.0625 to 3.75 in.).
- Type of pump (rod or tubing).
- Type of barrel (heavy, thin, or X type).
- Seating-assembly location (top or bottom).
- Type of seating assembly (cup or mechanical).
- Barrel length (ft).
- Nominal plunger length (in.).
- Length (in.) of upper extension (if required).
- Length (in.) of lower extension (if required).
Fig. 11.5 shows the API nomenclature for pumps covered by API Spec. 11AX. For example, a 1¼-in. bore-rod-type pump with a 10-ft heavy-walled barrel, a 2-ft upper extension, a 2-ft lower extension, a 4-ft plunger, and a bottom-cup-type seating assembly that will be used in 2 3/8-in. tubing would be designated as 20-125-RHBC-10-4-2-2.
Fig. 11.5—API Spec. 11AX description requirements for standardized pumps and the available options for the various components.
It is important to know that the users of API pumps need to provide, along with the pump nomenclature, the following ordering information: barrel and plunger material, plunger clearance (or fit tolerance), and valve (ball and seat) and fittings material. The materials normally available for each of these components also are now included in the latest edition of API Spec. 11AX.
Non-API and Specialty Pumps
The types of pumps, sizes, and component materials that are included in the API standards are based on the best industry practices that meet the widespread industry needs. While API standardizes the majority of pumps and components that are used in sucker-rod lift, there are special parts and pumps that have been developed by manufacturers to try to solve specific pumping problems. This specialty equipment should be considered when best industry practices and standardized components have proved unacceptable. However, the manufacturer of these components should create all parts to the same quality level required in API Spec. 11AX. Useful specialty pumps include the following:
- Casing pump for production without tubing.
- Pumps with two plungers that act in series to increase displacement.
- High-compression plunger assembly or pump for handling gas-interference problems.
- Three-tube pump for handling fines or solids.
- Pumps with a shorter barrel than normally recommended, so that the plunger completely wipes solids free of the barrel and prevents sticking.
Additionally, there are special pump components, such as valve rods, valves, and tubing drains, that are sometimes beneficial in situations in which the capabilities of normal API pumps and components have been exceeded. The manufacturer of special, non-API pumps and components should be contacted to determine the working capabilities and limitations of any of these specialty components. However, these items should be selected with care and used only after the best production effort has been thoroughly tested with standard components.
Materials SelectionThe most recent API Spec. 11AX was modified to add not only new sizes and types of pumps with new quality, inspection, and tolerance requirements, but also standardized, widely used pump-component materials. Table 11.3 presents the various material descriptions, their API identification symbol, surface condition, base core hardness, base material, and base-material minimum yield strength for plated barrels, as shown in API Table A of Spec. 11AX. Similar tables in Spec. 11AX (B through I) are incorporated for case-hardened barrels, nonhardened barrels, balls and seats, cages, pull tubes, valve rods, fittings, seating cups, spray-metal plungers, and plated plungers. These changes have incorporated the prior information in API RP 11AR and the Natl. Assn. of Corrosion Engineers (NACE) MR 01-76 for materials to be used in most production environments.
Allowable Setting DepthIn the early 1990s, an industry task group analyzed the stresses that react on a downhole rod pump. This was required to determine if there were recommended allowable loads that could be subjected to rod pumps of different types, sizes, and metallurgy. This group developed the burst, collapse, and axial-loading equations to determine these limits and the associated maximum recommended setting depth for sucker-rod lift pumps, published in API RP 11AR ; an example of the recommended setting depth of this standard is presented in Table 11.4. The depth limitation and stresses on the downhole pump barrel and components should be considered when selecting the size, type, and metallurgy for a downhole pump.
Slippage Past Plungers
The slippage or leakage past a plunger on a closely fitted sucker-rod pump is an important factor in properly designing and operating a well. The previous edition of the Petroleum Engineering Handbook discusses the main factors that affect leakage. Eq. 1 in Chap. 8 on sucker-rod pumps can be rewritten, combining constants, as the following equation:
in which Q = slippage or leakage loss, in.3/min; D = plunger diameter, in.; P = differential pressure across plunger, psi; C = diametrical clearance between plunger and barrel, in.; μ = absolute viscosity of fluid, cp; and Lp = plunger length, in.
The importance of plunger leakage is demonstrated in the example in the previous edition of the Handbook that shows for a 0.003-in. clearance, a 2¼-in.-diameter pump with a 48-in.-long plunger operating with a pressure differential of 2,000 psi at 15 strokes per minute (spm) and a 48-in. stroke length. Tight clearances (less than 0.003 in.) may cause producing problems, whereas loose clearances (greater than 0.008 in.) may result in excessive leakage by the pump. Good field-pump records are essential to make good pump recommendations.
Increasing the "compression ratio" of a plunger pump may reduce the effects of free gas and help prevent gas locking. The compression ratio is the volume of the pump chamber at the start of the downstroke divided by the volume at the end of the stroke. This ratio is fixed by the manufacturer on the basis of the design of the rod pump's components and the fit of the plunger to the pump barrel. Varying the sucker-rod pump components and close spacing will alter the compression ratio; however, some of these components are not standardized by the API Spec. 11AX. This can increase waste space in the pump, resulting in a decreased compression ratio. The importance of the compression ratio and associated waste space may prevent a new pump from being able to pump down a well. This work by McCafferty is further discussed in Hein, which also presents different pump manufacturers' normal compression ratios for similar pump types.
Selection of Subsurface Rod Pumps
Pumps for sucker-rod lifted wells should be selected on the basis of numerous variables that are provided by the well, the operating conditions, and the life of the pump. The main variables to consider are as follows:
- Well depth.
- Bottomhole temperature.
- Fluid viscosity.
- Amount and size of particulates in the produced fluids.
- Produced-fluids corrosivity.
- Required production rate vs. pump capacity.
- Fluid-specific gravity.
- Casing/tubing size.
- Well-completion type.
- Gas/liquid ratio (GLR).
- Pump-intake pressure vs. fluid bubblepoint.
- Spare/surplus pumps and components.
- New purchase and repair costs.
These variables influence the stresses on the pump, type of pump used, component metallurgy, pump size, internal-fit tolerance, and ability to handle solids/gas. Discussing these parameters with the pump manufacturer and local pump shop should help determine the proper pump to ensure acceptable pump life.
There are two aspects to consider when sizing the downhole pump for an installation. The first is that the pump capacity should be related to the well capacity. The pump displacement is determined on the basis of the pumping speed, unit stroke length, and plunger diameter. This general equation is
The stroke length should be the expected downhole stroke or plunger stroke (Sp) that is calculated from a sucker-rod string calculation or sizing computer program. However, the surface stroke length may be considered an approximation of the maximum capacity for a given pumping situation.
The recommended relationship of pump displacement to well capacity (WC), as discussed in Hein, is as follows:
Thus, for a well that produces 100 BFPD, the various pumping parameters should be selected to provide a pump displacement of between 118 and 154 BFPD. Because the pump displacement is greater than the well capacity, the system will require some type of well control to prevent constant operation and overpumping of the well. This increased capacity accommodates pump wear and loss of efficiency with time. As this occurs, system control should be adjusted to continue producing as required, without overpumping by running the pump more often. It should be considered that as the pump diameter increases, the efficiency of the system increases. However, this also increases the load on the rod string and the peak torque for the pumping unit. Thus, reasonable selection of these pumping parameters should be considered that results in extended run time.
The second aspect of pump sizing, once the pump diameter is selected, is ensuring that the downhole pump is properly built. The main component that needs to be sized is the barrel length, which should be long enough to accommodate the plunger length, the downhole stroke length, all fittings, and a rounding factor.
The minimum plunger length recommended is normally 3 ft. It is recommended that the length of the plunger is increased 1 ft/1,000 ft of well depth, up to a 6-ft maximum length. Plunger lengths longer than 6 ft have not shown to be an advantageous, while specialty pumps may have a plunger shorter than 3 ft.
When determining the barrel length, normally the maximum pumping-unit stroke length is considered to allow pump displacement to be increased with the existing downhole pump without pulling the downhole pumping equipment to change the capacity. However, this extra length and the pump-displacement option increase the price of the pump. Thus, the downhole Sp length should be considered the stroke measurement to use in the barrel-length calculation.
The types of fittings and their respective lengths depend on the type of pump being used. Normally, 12 to 18 in. covers the length range for various pump types.
The final factor in determining the barrel length is a rounding factor. Once the previous factors are added together, the length-of-barrel calculation is normally increased to the next available whole-foot standard length for a pump according to API Spec. 11AX. Using the surface stroke length vs. the downhole Sp length, and designating this length as the rounding factor, may provide sufficient barrel length to accommodate the spacing length some operators or pump shops suggest.
This spacing factor is normally a minimum of 24 in. for wells up to 4,000 ft deep, then increases 6 in. in length per 1,000 ft of increased well depth. These rules are recommended for all steel sucker-rod strings. When fiber-reinforced plastic (FRP) rods are used, additional increased spacing may be required because of the increased "stretch" or elongation of the rod string under the load. The FRP-rod manufacturer should have, or have access to, a sucker-rod-string design program that will estimate the increased plunger travel. This length then should be used in the barrel-length determination. Thus, for a 5,000-ft-deep well, with a required 74-in. surface stroke, a 48-in.-long plunger with a steel rod string and a designated 2 7/8 × 1½-in. RHB pump, the displacement length must be greater than 152 in. to permit adequate spacing. A standard 12-ft barrel with 1-ft top and bottom extension couplings should be considered.
Pump Operating Problems and Solutions
There are four common ways subsurface rod pumps are abused. These problems may also be applicable to other downhole pumps, and thus, these related solutions probably are applicable to other artificial-lift techniques. The four common abuses follow:
- Overpumping the well.
- Gas interference.
- Pump hitting up or down.
- Trash entering the pump.
Because the recommended pump-displacement design is for the pump to have greater capacity than the well, an overpumping condition may occur if the well is not properly controlled. An overpumping condition is indicated when there is a fluid pound more than one-quarter of the way down on the downstroke because of insufficient fluid in the well to charge or fill the downhole pump. This condition may be seen on the surface if the pound is very severe, but the best way to detect this is with the use of a dynamometer. Other indications of overpumping are if the pump volumetric efficiency is less than 70% or if a downhole fluid-level survey shows that the normal operating fluid level is at or very near the pump intake. Overpumping may cause mechanical damage to the pump or cause damage uphole to the rod/tubing because of increased buckling and wear. Properly setting a well controller will help reduce severe overpumping.
Indications of gas interference include low volumetric efficiency, while the fluid-level survey shows apparent, adequate pump submergence and a polish rod that is excessively hot to the touch. A dynamometer survey, when combined with the precalculated well loads for the applicable design conditions, may indicate gas pound, gas lock, or inconsistency with the assumed conditions. The gas-interference condition may be remedied by increasing the pump compression ratio, if possible. This may be as simple as respacing the pump as the fluid level decreases in the well annuli or changing the stroke length for the pump downhole, or it may require pulling the pump and altering its design. The compression ratio of the replacement pump should be determined to ensure adequate lift capabilities. Additionally, a pump with tighter fit tolerance/waste space, smaller pump diameter, increased stroke length, adequate downhole separation, and properly designed pump gas anchor should be considered along with properly placing the pump intake above or below the perforations, as previously discussed. Finally, if these normal solutions do not resolve the problem, then special pumps or specialty components may be considered.
A pump component hitting on the up- or downstroke is indicated by an instantaneous load change and can be shown with a load-capable dynamometer. This condition normally occurs because of inadequate pump spacing as the fluid level pumps down or because the pump has inadequate compression ratio/excessive waste space for the seating depth for the designed pumping parameters. While severely "tapping," or "tagging," the pump may be heard, felt, or seen, the smashed pump components obtained during a pump teardown will show the damage this condition causes. This condition may also be magnified for tubing that does not have an anchor, or if the anchor is not properly set. Other conditions that may cause this problem include if the pump-intake piping is plugged or not properly designed, if the pump has inadequate compression ratio, if the polished-rod clamp is not sufficiently tightened, and/or if the pump barrel is not properly sized.
The last normal operating problem is caused by solids entering the pump. There are many reasons for these particulates. The particulates may be caused by well conditions such as producing the fracturing sand back into the wellbore, very fine powder from the formation, iron sulfide scale from the downhole equipment because of inadequate corrosion inhibition, iron sulfide or other scales from the formation because of incompatible fluids, or from overpumping the well. Solutions include using different types of pumps designed to handle fines and solids, such as three-tube pumps or soft-packed plungers, and using harder materials or coatings for the pump components. Filters or downhole, wire-wrapped screens have been used with limited success until they plug. In the past, tighter fit tolerances (< 0.003 in.) for the plunger-barrel annuli have been considered; however, recent work done in both the laboratory and the field, has shown the benefit of increasing these tolerances to greater than 0.005 in. when solids are a problem. This work has resulted in the variable-slippage pump that would be useful for conditions in which solids are present in the produced fluids and gas interference is also a problem.
Pump Shop, Repair, and Audit
The pump manufacturer typically machines or obtains subcontract pump components for future assembly of the pump by a pump shop. The shop, the knowledge of the design, selection of pump types, and associated component metallurgies become critical to long well life and a decreased failure frequency. API RP 11AR provides useful information on pump types, component and metallurgy selection, pump-setting-depth calculation, and pump assembly/teardown.
While the pump manufacturers usually produce their pump components with an acceptable quality program (such as ISO 9001 or API Spec. Q1), most pump shops are not covered under these rigorous plans. Thus, it becomes critical to have the pump shop and its employees audited by qualified personnel to ensure that training, workmanship, safety, and environmental considerations are adequate. On the basis of many shop audits, assembly and teardown observations, requirements and recommendations in API standards, and performance quality requirements, a checklist that should be used as a first step in obtaining an acceptable pump shop has been developed and published. Once the audit is performed and the checklist completed, the findings should be discussed with the appropriate pump-shop personnel and a time line developed detailing when changes to resolve any problem areas will be made.
11.4.1 Steel Sucker Rods API Spec. 11B provides the industry requirements for sucker rods and some related sucker-rod lift equipment. The three main grades of steel rods follow:
- Grade C rods that have minimum and maximum tensile strengths of 90,000 and 115,000 psi, respectively.
- Grade K rods that have a minimum tensile strength of 90,000 psi and a maximum strength of 115,000 psi. These rods are made with 1.65 to 2.00% nickel and are, therefore, more expensive than Grade C rods, but may have improved corrosion-related properties.
- Grade D rods that have a minimum tensile strength of 115,000 psi and a maximum strength of 140,000 psi. Three types of this grade are covered by Spec. 11B: plain-carbon, alloy, and special-alloy steels.
Spec. 11B allows for rod lengths of 25 or 30 ft and pony rods in six lengths (i.e., 20, 44, 68, 92, 116, and 140 in. measured from contact face of pin shoulder to contact face of pin shoulder). The acceptable rod diameter goes from 5/8 to 1 1/8 in. in 1/8-in. increments. The most common rods in use will meet API specifications and will probably be in 25-ft lengths. The most important selection requirement is that the pulling rig can accommodate single-, double-, or triple-length rod segments.
The API does not specify the minimum yield strength for sucker rods. Where the yield strength of a rod string is necessary in calculations, it is recommended that if the manufacturer is not known, a minimum yield of 60,000 psi for Grade C and K and of 100,000 psi for Grade D should be used. If the manufacturer and rod type are known, the actual yield-strength values may be used. For good operating practices, the minimum yield strength should not be exceeded.
API RP 11BR provides industry recommendations on the selection and use of API-grade rods.
Pony rods are sucker rods shorter than 25 ft, and they vary in length. They are most commonly placed adjacent to the polished rod at the top of the rod string, on top of the downhole pump for handling purposes, and on top of the polished rod with appropriate couplings to prevent the string from falling downhole if the polished-rod clamp slips. Old pony rods normally should not be used in the load-carrying part of a new rod strings. Thus, when placing the rod string with new suckers, new pony rods should be used.
FRP Sucker Rods
FRP sucker rods may be used instead of metal under certain conditions. These rods are normally made from protruded fiberglass. They also are standardized in size and performance by API Spec. 11B. Reviewing this standard shows that temperature, load reversals, and fatigue life have a bigger effect on FRP rods than on steel rods. It is important to keep the following in mind when screening a well for FRP-rod use:
- FRP-rod bodies will not corrode, but the rest of the steel components, including the fiberglass pin connectors and couplings, the steel rods making up the rest of the string, the pump, tubing, casing, flowlines, etc., still have to be protected if producing a corrosive fluid. Thus, fiberglass rods should not be used alone to prevent rod-string corrosion or system failures or to eliminate the need for an effective corrosion-inhibition program.
- FRP rods should be considered when the pumping-unit gear-reducer torque or structure rating exceed design limitation and need to be decreased. Reducing the weight of the sucker-rod string reduces the torque measured at the polished rod. However, if the well is expected to produce long term, it may be more cost effective to upsize the pumping unit.
- It should be determined if it will be possible to stroke the subsurface pump plunger because of the increased elasticity and effect on Sp .
- If the well deviation is very large at any point, the increased friction may cause buckling and compressive stresses on the sucker rods. Increased buckling is very damaging to FRP rods; thus, these probably should not be run in deviated wells.
- Allowing fluid or gas pounding may produce damaging compressive forces in the FRP rods; thus, maximum drawdown is not possible.
Currently, there is no recognized formula for calculating overtravel when a mixed FRP and steel rod string is used. An attempt was made by an API task group to try modifying API RP 11L to include a FRP-rod-string analysis, but this was not accepted by the industry. A study of several FRP string-design analyses indicate that rod-string overtravel may be approximately equal to the following:
where S = stroke length, in.; N = pumping speed, spm; and LPSD = seating nipple/pump depth, ft. This overtravel approximately equals twice the expected value when using steel sucker-rod strings.
Non-API Sucker Rods
Non-API sucker rods generally fall into two groups: one contains rods with a higher strength than API Grade D, and the other contains rods made of alloys that are less susceptible to corrosion or that have received a special heat treatment.
The high-strength group is generally harder and higher strength than Grade D and may be more susceptible to hydrogen embrittlement and notch effects that may then decrease run life.
Those rods that have a special heat treatment or are made of special alloys are normally premium-priced items. Thus, a full economic analysis should be conducted and good operating records obtained to determine if use of these rods is cost effective.
Flexible Strand. Approximately 40 years ago, a top steel manufacturer experimented with the use of plastic-coated wire cable instead of sucker rods. This cable was a continuous strand that required special pulling equipment. Sufficient sinker bars or a special pull-down pump had to be used to keep any compressive force from acting on the strand. The connectors used at the pump or at the top of the sinker bars were the weakest portion of the flexible strand. If any of the strands furnished the weight that was required to help open the traveling valve, the strands immediately above the sinker bars failed in short order because of the compressive forces. This type of rod string was less expensive than a normal API steel string and was found useful for unloading gas wells. The biggest disadvantages that restricted the use of this type of string were lack of service-company support and the inability to make field repairs.
Continuous Solid Rod (COROD*). The advantage of this rod is its ability to pull the entire rod string in one piece with a special pulling unit. These rods are available in either round or elliptical configurations and vary in size from 12/16- to 18/16-in. diameter. The disadvantages include the need for a special wheeled pulling rig, and the two different pulling units are required to service the well if the tubing has to be pulled. There is some concern that the COROD's heat treatment is not consistent throughout its length. This is especially problematic if field welds are made and the rods are used in an inadequately protected corrosive environment.
A continuous strand of composite materials, called "ribbon rods," was developed and field tested. This type of special rod contained carbon composite with a polymer wrap. Despite having high strength and a small cross-sectional area, it was expensive and ran into field support problems similar to those of flexible strands and CORODs.
'"Electra" Sucker Rods. Another type of non-API sucker rod is the Electra (EL)** rod. These currently are available only in 3/4-, 7/8-, and 1-in. diameters. They should be selected for wells in which operating stresses do not exceed 50,000 psi. These rods have a special heat treatment that should put the surface in a compressive set. Thus, they could be used in a hydrogen sulfide (H2S) environment in which the strength of Grade C rods is exceeded. These rods have been effectively used to produce approximately 150 BFPD from a depth of approximately 14,500 ft.
High-Strength, Low-Alloy Rods. A number of manufacturers have developed higher-strength steel rods to compete with other specialty rods. These rods take advantage of the newer alloys and heat-treating procedures currently available and are based on American Iron and Steel Inst. (AISI) 8630- or 4130-type steels, which have high tensile strengths. The tensile strength is generally greater than 140,000 psi, while the yield strength is generally greater than 100,000 psi; therefore, these rods could not be classified as API Grade D. The fine-grain heat treatment done on these alloys theoretically should provide increased fatigue life. However, this rod type may be more notch-sensitive and may require better handling and corrosion protection than normal API-type rods.
As with any specialty equipment, good field testing and records for several years in which good handling and operating practices were followed are required to prove the benefit for any of these non-API rods.
Criteria for Rod-String Design
Rod Stress. In a noncorrosive environment, the endurance limit of steel is primarily determined by the maximum stress, the range of stresses, and the number of stress reversals. This is often illustrated by the use of a Goodman diagram, as discussed in API RP 11BR. Derating, or service, factors also are discussed to allow potential decreasing of the load range for different service/corrosive environments. If the environment is corrosive and not properly treated, the sucker rods and their associated downhole equipment life is minimal. In such cases, corrosion-fatigue failures occur frequently in the rod string.
Effectively inhibited systems may be considered noncorrosive, which would limit the surface pitting of the steel rods or components. However, in the presence of H2S and a corrosive environment, steel may become susceptible to hydrogen embrittlement/sulfide-stress cracking. Steels that have a Rockwell C hardness greater than ≈ 23 (Brinell hardness number 237) are susceptible to embrittlement. The harder the steel is, the more susceptible it becomes. API Grade C sucker rods normally have a Rockwell C hardness < 23, while API Grade D sucker rods normally have a Rockwell C hardness > 23. Thus, API Grade D rods should be used with caution in the presence of hydrogen sulfide. Chemical inhibition may not prevent embrittlement. This results in a significantly decreased run life.
Stress raisers cause areas of concentrated stresses and may be caused by a number of things. Corrosion pits are one type of stress raiser. Stress raisers may be notches caused by improper handling, tool cuts, bending, and subsequent cold straightening, for example, and may also result from the manner in which the threads are formed on the rod pin (i.e., cutting vs. the now-required cold rolling). Corrosion pits may have rounded or notched shape; notch-shaped pits are more serious and are more likely to occur in Grade D rods than in Grade C rods.
API RP 11BR recommends using the modified Goodman diagram for determining the allowable stress on API steel-grade sucker rods, while API Spec. 11B covers FRP rods. Manufacturers of non-API rods should specify the rod's allowable stress. An allowable load or stress curve should be developed to discern during the design of a rod string if it is overloaded, and adjustments should be made to prevent this. Recent discussions have promoted a hyperbolic relationship for allowable load using the Gerber parabola, rather than a straightline relationship. This loading criterion, coupled with cleaner steels and better-quality sucker-rod manufacturing, should enable higher allowable loads to be applied to the rod strings, provided that good sucker-rod handling practices are followed. Rod strings that are considered "overloaded" by more than 20%, according to the straightline method, have been successfully run in the Permian Basin fields in the U.S.A. and provided adequate run time. Additionally, RP 11L, discusses the need to reduce the allowable load or stress on used rods. Recommendations are presented for derating based on the class of the inspected rod, according to the inspection-criteria classes in API RP 11BR.
Rod-String Selection. The primary factors affecting the selection and sizing of rods and the rod system are as follows:
- Size of pump and tubing.
- Liquid viscosity and pourpoint.
- Kind of corrosion [e.g., H2S, carbon dioxide (CO2), or saltwater].
- Conditions for unseating the downhole pump.
- Pump setting depth.
- Production rate.
- Sand, paraffin, salt crystals, scale, foam, and GLR.
These factors should be considered when manual (according to API RP 11L) or computer design calculations are performed to size the rod string and the related production equipment for a specific well.
Sucker-rod strings may be composed of a single size or may be tapered, typically to include rods of two and three sizes. Using four or more sizes of rods in a taper is not normally recommended. The primary factor determining the proportion of each size of rod in the rod string is the size of the pump. However, typically only one grade of rod is used in the string to avoid mixing during running and pulling operations.
API RP 11L contains recommended rod-string design data. The first column of Table I in this reference contains the rod-string size designation. The first number in the column refers to the largest rod size in the string, while the second number refers to the smallest rod size in the string, both representing the size in eighths of an inch. An example rod number of 76 is a two-way taper of 7/8- and 6/8 -in. rods. Rod number 86 is a three-way taper of 8/8 -, 7/8-, and 6∕8 -in. rods.
Rod strings should be designed to enable the operator to unseat the pump without yielding any rod in the rod string. The diameter of the pump plunger determines the fluid load lifted during the pumping cycle. However, the ID of the seating nipple determines the fluid load that must be lifted to unseat the pump. Friction in the pump holddown plus sediments in the pump-tubing annulus increases the required pump-unseating force. However, a high tubing-casing-annulus fluid level decreases the load on the rod string when attempting to unseat a pump. Normally, the pulling-rig weight indicators are not accurate enough to use as the only tool to prohibit yielding the sucker rods. The rod string's stretch in Table 4.1, Column 4, of API RP 11L, gives elastic constants (Er) for sucker rods that can be used to indicate rod load.
The top rod in the bottom section normally has the highest stress in the string because it has the smallest cross-sectional area. This is because it has to support the weight of the rest of the small-diameter rod load, the pump and the very large fluid load on the gross seating nipple area. The weak point in the string is this rod. A free-body diagram can be used to determine the loads acting on this rod; an allowable unseating load or stretch can then be determined so that the rods are not yielded or damaged when trying to unseat the pump.
To Taper or Not To Taper a Rod String
Tapered rod strings that use different segments of different-sized rods are commonly used to save unnecessary weight and to distribute the loading on long strings of rods used in deep wells. The proper design will decrease the stress on the rods above the bottom section. This allows pumps to be run deeper than would be possible if just one size of rod was run. Tapered rod strings can be operated at a higher pumping speed (N) than straight rod strings. This may reduce the required pumping-unit gearbox size and increase rod stretch because stretch is proportional to rod-string weight. Thus, more production may be possible from the well with a tapered string than a straight string using the same-diameter pump.
Ideally, a rod string should be a continuous taper from top to bottom. This is impractical, not only because of the manufacturing difficulties involved, but also because the lower rods must have sufficient stiffness to support the entire string in the tubing if failures occur high up in the string. For this reason, 75 to 85 strings are not normally recommended because, if the rod string parts high in the well, close to the surface, the 5/8-in. rods may be permanently damaged when the upper rods fall on them. Coupled sucker rods come in diameter variations of 1/8 in. With the introduction of the continuous sucker rod, the opportunity for a greater number of tapers is possible because these rods may be manufactured in size variations of 1/16 in. or even smaller.
The primary factor in determining the proportion of each size of rod in the rod string is the size of the pump. Columns 6 through 11 in Table D.1 of API RP 11L contain the percentages of the various sizes to be placed in a tapered rod string with various pump sizes. Before 1977, percentages were calculated so that the unit stress on the top rod of each section from the weight of the rods in air plus the weight of the produced fluids on the gross plunger area is equal. This is calculated as a static load. Work done by API and Shell in 1977 resulted in the percentages shown in API RP 11L. This work used the dynamic effects on the rod's upstroke and downstroke, along with assumed pumping speeds for varying stroke lengths. Currently, most operators and rod manufacturers have proprietary rod-string design programs that include these data.
One of the earliest means used for designing tapered sucker-rod strings is in the Sucker Rod Handbook. This design is based upon equal stress in the top of each size of rod, assuming a static condition and pumping water (specific gravity = 1.0) with the well pumped off. Buoyancy of the rod string is not taken into account. The recommendations in API RP 11L3 are based on the same assumptions. However, continued work suggested adopting a "modified-stress" approach in which the stress from the dynamic loads at the top of each size of rod is equalized. Computer programs are available to perform the calculations on this complex process of assessing stress for various rod-string designs.
Rod CouplingsAPI Spec. 11B contains requirements for the rod couplings, as well as the rods, and recommends minimum tubing sizes. The current edition provides for two classes of couplings: Class T (through hardened coupling) has a Rockwell C hardness range of minimum 16 and maximum 23, and Class SM (surface hardened) has a minimum Rockwell C surface hardness of 50. This hardness is normally accomplished by the spray-metal process. Care should be taken when recommending the SM couplings, even though they have longer wear life than T couplings. Because of the increased hardness and lower coefficient of friction, if properly surface treated, coupling-on-tubing wear is transferred from the rods—which are easy and less expensive to replace—to the softer tubing, which is more expensive to replace. Thus, while the SM couplings help to increase rod-string life, the tubing life may be decreased. API Spec. 11B also standardizes "full-sized" coupling in both grades and a "slimhole" coupling in Class T. Tables 4.1 and 4.2 from API Spec. 11B shows recommendations for the minimum tubing sizes for the various couplings.
Slimhole couplings for 5/8- to 1-in. rods can be run and fished in one-size-smaller tubing than the respective full-sized coupling. This enables operators to run 1-in. rods in 2 7/8-in.-OD tubing and 7/8-in. rods in 2 3/8-in.-OD tubing. This coupling type, however, decreases the coupling area available for supporting the pumping loads. Thus, slimhole couplings are not as strong as the full size. Original work by Gipson and Swaim recommended derating these couplings on the basis of the assumption that the 1-in. slimhole coupling has an acceptable minimum decreased area. Further work by Hermanson using the area relationships and allowable strength of the different grades of steel rods resulted in different derating factors, shown in Table 11.5. Additionally, these have been accepted by the industry and included in API RP 11BR. Note that the use of 7/8-in. slimhole couplings results in the highest derating factor for all rod strengths and sizes.
Well equipment, including sucker rods, must be in good working condition. The sucker-rod string is often highly stressed and usually fails because of the repeated load reversals. Corrosion, scale, and paraffin deposits may accelerate such failures. Tubing and rods will wear because of the reciprocating movement in the well caused by pounding fluid, buckling because of unanchored tubing, and/or bad wellbore deviation that allows contact.
Sucker-rod strings are lifting a great deal of weight every cycle. They are under stress on both the downstroke and the upstroke. Combining this with the normally corrosive environmental conditions of water, H2S, CO2, etc. may mean that one of the greatest expenses of a producing beam-pump system is replacing the sucker rods. Carrying out the various procedures described in this section can greatly reduce operating costs and make production more efficient and economical.
Care and Handling of Sucker Rods. Proper running, handling, and makeup procedures should be followed to secure maximum service from a rod string. API RP 11BR contains the practices recommended by the industry.
Torque measurement has been discredited as a sucker-rod-connection makeup method. When the threads are properly lubricated, an estimated 10% of the applied torque turns the coupling relative to the pin, and 90% of the torque is consumed by friction. Any variation in lubricants or in the surface finish of the threads or mating surfaces drastically changes these percentages, indicating that torque could never be a precision makeup method for sucker rods.
API RP 11BR recommends circumferential displacement (CD) for making up sucker-rod joints, and it should also be used for calibrating power tongs. To make up a sucker-rod joint using CD, the pin and coupling threads should be cleaned and lubricated with a lubricant that has passed the NACE MR-01-74 screening test. This test states that an acceptable lubricant will allow the lubricated pin to be made up hand tight, then fully made up and broken out 10 times without galling the threads. A hand-tight position is attained when full shoulder abutment is made and a 0.002-in.-thick feeler gauge cannot enter into this interface between the rod and coupling face. The coupling should then be turned by the amount specified in API RP11BR or by the rod manufacturer, relative to the pin. The manufacturer of specialty or non-API rods should be consulted for their recommended CD values and makeup procedures.
Rod-String Equipment Failure. The downhole production strings may fail for a variety of reasons, some of which have been discussed previously. Steward and Moore discuss reasons for common sucker-rod string failures and provide discussion and pictures of the failures. Additionally, Hermanson provides discussion and photographs of different rod failures. The following is a summary of the normal rod-string equipment and typical reasons for failure:
- Polished rods.
- Not in center of tee throughout pumping cycle.
- Smaller than recommended by API.
- Top of carrier bar not horizontal.
- Crooked—not vertical—wellhead.
- Crooked hole near surface, with pony rods below the polished rod.
- Excessive heat.
- No lubrication.
- Packing too tight.
- Pony rods (rod subs).
- Old subs used with new rod string.
- Improper API-grade rod.
- Sub directly below polished rod.
- Rod couplings (boxes).
- Slimhole couplings used.
- Hammered-on boxes.
- Insufficient circumferential displacement.
- Dirty or improperly cleaned threads.
- Improper or no lubricant (should be a properly screened inhibitor, not tubing or drillpipe dope).
- End face not perpendicular to the threads.
- Oxygen in system.
- Couplings made from free-machining steels.
- Rod pins.
- Old-style, nonundercut pins.
- Incorrect circumferential displacement.
- Box and pin not made up, but broken out and remade on new C and K rods.
- Box shoulder and pin shoulder not parallel.
- Rod upsets.
- Worn elevators.
- Rod bent while tailing out or in.
- Rods corkscrewed above the pump during normal pumping.
- Rods corkscrewed after parting.
- Manufacturer's marks.
- Running too fast in the hole.
- Rod body.
- Inadequate/ineffective corrosion inhibition.
- Hydrogen embrittlement.
- Service time exceeds fatigue life.
- Rough surface.
- Yield strength exceeded while attempting to unseat pump.
- Defective material.
- Oxygen allowed in the pumping system.
- Valve rod (stationary barrel pump).
- Pump not centralized in tubing.
- Improper material.
- Plunger too short and pump not centralized.
- Crooked hole at pump setting depth.
- Pounding fluid.
- Pull tube (traveling barrel pump).
- Pump not centralized in tubing.
- Pull tube buckling on downstroke.
- Improper material.
- Pump set too deep for pull-tube length.
- Pounding fluid.
Replacing a rod string one rod at a time is not normally a good operating practice; thus, the economic life of a rod string needs to be considered if rods start to fail. Typically, the rod-string section will be replaced after two or three failures, while the entire rod string may be replaced after three or four failures. However, the reasons for failures need to be investigated and the root cause for this failure must be determined to extend the rod life in the future.
An SPE paper by Powers considers the factors that enter into the decision about when to replace the entire rod string after sustaining the calculated number of failures. Usually, wells of the same type in a field can be grouped together and the necessary calculations do not have to be performed for each well. Sufficient calculations need to be done to assess the economic impact for all wells in a field.
* COROD is a product of Weatherford Intl. Ltd., Houston.**
Miscellaneous Subsurface Equipment
TubingThe chapter on tubing selection, design, and installation from this Handbook provides detailed information on the design, selection, and use of tubing for production wells. As related to most sucker-rod-lifted wells, the standard weight of external-upset-end, API tubing should be used because of the increased wall thickness in the threaded ends. Thus, if there is rod coupling-on-tubing wear, more life and fewer leaks will be realized than if nonupset API tubing is used. Using API Grade J55 tubing, consider full-body normalizing after upsetting to prevent "ringworm corrosion" in the heat-affected upset region when the tubing is placed in corrosive (H2S or CO2) service. If the production application is noncorrosive, then this extra heat treatment may not be required.
Tables 4.1 and 4.2 from API Spec. 11B include minimum tubing size for each size of full-sized and slimhole rod couplings. There should be sufficient clearance between the tubing and the rod box for fishing tools.
The yield strength of the tubing must be sufficient to support the weight of the tubing in air, the weight of the rods and of the fluid in the tubing, plus an overpull allowance that will allow the tubing to be pulled. Normally, API Grade J55 is acceptable for most rod-pumped wells to a depth of approximately 9,500 ft. However, with greater well depths and higher production rates, API Grade N80 or L80 (if H2S is present) and, in some cases, P110 should be considered.
It is recommended that API tubulars be drifted to ensure equipment can be run without problems.
Thread dope must be used on API tubing threads to keep the joints from leaking, but it does not have an infinite life. If collar- or tubing-connection leaks begin to appear in tubing strings, it may be necessary to remove all collars (if applicable), clean the threads on the tubing and the collar or upset connection, and apply new thread dope. Additionally, tubing that has been in storage should at least be visually inspected, and the threads cleaned and freshly doped, following API recommendations, before running.
Most wells will be able to use normal torque makeup requirements for tubing. A guideline for appropriate makeup of oil-country tubular goods is found in API RP 5 C1. This RP also includes care and handling along with running casing and tubing information.
Hydraulic testing of tubulars in the well will determine only whether, under that circumstance, the tubing and couplings are leak free. Once the well is put back on pump, rod-on-tubing wear may reduce the wall thickness, causing a split. Additionally, hydrotesting itself may provide sufficient pressure to fail a worn tubular that may have had acceptable pressure retention to handle the pumping pressures. Thus, if tubing wear is a problem, downhole tubing-caliper surveys or surface tubular inspection should be done to separate unacceptably worn tubing before it leaks. Fig. 11.6 presents an example of a downhole tubing-caliper survey. It should be noted that the major wear is approximately midway between rod couplings because of rod buckling from pounding fluid. The chart also shows that there was wear caused by the couplings themselves contacting and wearing the tubing.
New developments have been made in using internally plastic-lined tubing in rod-pumped wells. Such tubing has been beneficial in preventing erosion at the pump discharge and/or wear along the inside of the tubing. One west Texas operator dramatically reduced the field failure frequency from 0.42 to less than 0.25 in the Howard Glasscock field by running full and partial strings and, in many cases, just a few joints of this poly-lined tubing on the bottom of the tubing string. Monitoring of these lined tubing joints should continue to ensure that the liner does not wear or degrade with time.
The failure frequency is a dimensionless number found by dividing the total downhole well failures by the total number of producing wells in a field. This failure frequency can be further described by dividing the number of sucker-rod, tubing, or pump failures in a year by the total number of sucker-rod-lifted wells to determine which equipment is causing the most failures in the field. Similar calculations can be done for other lift methods that are used in the field.
Tubing-Anchor Catchers (TACs)Tubing anchors are used to prevent movement of the tubing during the pumping cycle. Fig. 11.7 shows an example of the recommended mechanical-type TAC for rod-pumped wells. During pump operation, part of the fluid load is transferred from the tubing to the sucker rods, alternately. This causes the tubing to elongate on the downstroke when it supports the fluid load and to shorten when the rods carry the fluid load on the upstroke. This action shortens the effective plunger stroke and decreases the pump displacement. This load transfer also causes helical buckling in the bottom portion of the tubing string, which, in turn, causes additional rod-on-tubing wear. The recommended TAC has two-way slips; these prevent parted tubing from falling in addition to preventing movement during the pumping cycle.
Tubing anchors are normally placed within 30 to 100 ft above the pump's seating nipple. The tubing is set in the surface hanger with tension equal to the sum of the tensions required to overcome the stretch because of load transfer, helical buckling, the anticipated temperature change between producing the shut-in conditions, and the change in fluid level. A calculation procedure from the manufacturer should be followed to properly set the TAC "total stretch," rather than pounds of pull from the rig. Further consideration should be given for adequate settings, if the downhole pump diameter exceeds the tubing diameter, as in the case of oversized tubing pumps (sometimes called casing pumps). When this occurs, the normal applied stretch or load for the tubing has shown to be inadequate, requiring increased stretch-setting inches.
This equipment can be difficult to remove; thus, care should be taken using a TAC in wells having scale, heavy paraffin, sand production, and/or bad casing. The TAC release method should be considered before this equipment is installed.
Several of the tubing anchors available have shear pins to release the slips if the normal releasing mechanism fails. Varying the material type and number of shear pins can vary the amount of necessary pull; this is called the "shear-out value." The tubing must have sufficient yield strength to support the weight of the tubing in air, the weight of the rods, and the weight of the fluid in the tubing as well as to shear the pins left in the tubing anchor. These factors will limit the pumping depth to which a TAC can be used. However, the running depth can be increased with stronger tubing and/or tapered tubing strings and with the required minimum strength and number of shear pins. Care should be used to ensure that the design shear out or production loads do not exceed the tubing-grade yield strength. If this possibility exists, the tubing should be cut rather than pulled apart.
Tubing rotators may be used to spread tubing wear because of rods and/or rod couplings around the entire diameter instead of being concentrated in one spot. They may be used in conjunction with rod rotators to even out the wear on both the tubing and rod coupling.
Tubing rotators come in more than one size. The manufacturer should be consulted when selecting these items to ensure the rotators purchased are sufficiently strong for the particular job. In most cases, the use of a TAC, coupled with rod centralizer and possibly a rod rotator, will prevent sufficient wear such that a tubing rotator is not required.
A sinker (or heavy-weight) bar is normally a special steel bar or large-diameter sucker rod placed directly above the downhole pump. Such bars may be used polished rods or a rod specifically standardized by API Spec. 11B.
During the pumping cycle, these bars help to open the traveling valve because a portion of the pressure required to open the valve on the downstroke must be obtained from the weight of the sucker-rod string pushing down on the top of the plunger. This places the lower portion of the rod string in reduced tension. Rod buckling will result unless properly sized and centralized sinker bars are used immediately above the pump to provide the additional needed weight. Sucker-rod buckling will cause excessive rod- and/or coupling-on-tubing wear above the pump. The buckling at the bottom of the rod string also may cause premature valve-rod or pull-tube failures. Overall, there are a number of advantages for using sinker bars in a sucker-rod string, which may include the following:
- Keeps tension on the sucker-rod string.
- Increases the minimum polished-rod load.
- Decreases polished-rod horsepower (HP).
- Decreases low tubing leaks.
- Decreases valve-rod or pull-tube pump failures if caused by buckling or bending.
- Increased production.
- Overall decrease in operating costs.
There also are disadvantages from using sinker bars, including the following:
- Creates added mechanical problems when the production equipment is allowed to pound fluid more than one-quarter of the way down on the downstroke.
- Increases operating expense if purpose-manufactured rods are purchased.
- Inadequate coupling makeup and pounding fluid can cause the connection to unscrew, if polished rods are used.
The theoretical sinker-bar weight required in a rod string depends on the specific gravity of the produced fluids, the size and type of downhole pump, the associated valve-seat contact area, and the depth of the well. There are differing thoughts on the minimum amount of sinker bars required. Some operating companies and sinker-bar manufacturers use a weight equal to the buoyant weight of the rod string in the produced fluid. Others use only 20% of the well depth or no sinker bars—only a few sucker-rod centralizers or guides near the bottom. Some operating companies use a sinker-bar factor (SBF) for the various types of pumps. Gipson and Swaim developed the SBF for stationary barrel pumps in the "Beam Pumping Fundamentals" (April 1969) and published them. Traveling-barrel pumps normally have a traveling valve one size larger than stationary barrel pumps; thus, these SBFs need to be increased.
The SBF process is to determine the theoretical weight of sinker bars in the produced fluids. Then, 20% of this theoretical weight is the recommended starting point for the actual weight or length of sinker bars used to replace the lowest rods in a rod string. This was recommended because sinker bars act dynamically to help valve action and to help keep the rods in tension. Once sinker bars are run, an optimization to increase the number of bars or weight can be conducted. However, there is a minimum point of benefit at which adding more sinker bars will not provide the useful dynamic effects. When this occurs, the extra bars or weight will be detrimental to rod-string loading.
An SBF summary for the theoretical weight for the various-diameter stationary and traveling barrel pumps is presented in Table 11.6. With these values, the recommended starting sinker-bar weight is as follows:
The resulting sinker-bar weight to install is as follows:
where LPSD = seating nipple depth, ft, and G = specific gravity of the combined fluid in the tubing.
Sucker-rod centralizers also may be called paraffin scrapers or rod guides. They keep the rods and couplings away from the tubing to decrease wear. However, special mechanical paraffin scrapers have been developed to also aid in keeping paraffin off the tubing and most of the sucker-rod length.
Rod centralizers with full-bore-fluted centralizers should be placed on or between the pump-handling pony rod, the sinker bars used above the pump, and the first two sucker rods above the sinker bars. Rod centralizers in these locations help stabilize the pump and valve rod and prevent valve-rod bending or breakage. When a tubing anchor is not used, rod centralizers will reduce tubing wear because of tubing helical buckling on the upstroke. Rod centralizers also may be used in crooked holes in which there are areas of concentrated tubing wear.
Sucker-Rod-Guide PlacementWhen setting rod guides, it is necessary to determine the correct spacing when the tubing anchor is set several hundred feet above the seating nipple or when a TAC is not run. It is recommended as a starting point to use the Lubinski curve to determine guide spacing; Fig. 11.8 provides the minimum guide-spacing curves for 2- and 2½-in. tubing.
The formulas for determining the distance that unanchored tubing will buckle above the seating nipple are as follows:
where Fo = 0.34 × G × D2 × H, which is the fluid load on the gross plunger area, G = specific gravity of the mixed fluid in the tubing string, D = pump-plunger diameter, and H = pump-seating depth in ft.
As an example problem, solve the following:
Given: tubing = 2 7/8-in. OD API, D = 1.50 in. (pump plunger diameter), L = H = 8,000 ft (pump-seating-nipple depth and assumed pumped-off fluid level), and G = 1.03 (specific gravity of the liquid in the tubing). A TAC is to be set at 7,450 ft, which is 15 ft above the top casing perforation.
Find: (a) the buckling distance and (b) the recommended spacing for sucker-rod guides.
1. buckling distance = Fo / 5.7 = [0.34 × 1.03 × (1.5) × 8,000] / 5.7 = 6,304 / 5.7 = 1,106 ft.
Fig. 11.8 indicates that when the neutral point is 1,106 ft above the seating nipple, the first guides should be approximately 15 ft apart, or approximately two guides are recommended per 25-ft-long sucker rod in 2 7/8-in. OD.
In summary, there will be 8,000 – 7,450 = 550 ft from the seating nipple to the anchor. The anchor will be 1,106 – 550 = 556 ft below the neutral point. Fig. 11.8 indicates that guides should not be less than 25 ft apart until approximately 380 ft below the neutral point; therefore, it is recommended that two guides be placed on each 25-ft-long sucker rod, between the seating nipple at 8,000 ft and the TAC at 7,450 ft. This is the minimum number of guides per rod.
If continued rod and/or coupling-on-tubing wear is a problem, more centralizers should be considered. Wellbore deviation is one of the biggest problems for sucker-rod-lifted wells. If the deviation is 0 to 3°/100 ft, there should be no pumping problem. A deviation of 3 to 5°/100 ft is a bearable problem, and it usually can be handled by properly locating the rod guides. A deviation greater than 5°/100 ft is a definite problem. An increased number of guides per rod, tubing anchors, and/or special roller rod guides may be necessary within the local deviation region.
Rod-Centralizer Types and Materials
There are two main types of sucker-rod centralizers: field installable or molded on. The field-installable guides can be hammered on, twisted on, or (with two pieces) slid together on the rod. Usually, these field-installable guides do not grip the rod area very well; thus, they do not stay where they are required. However, guide manufacturers continue to develop these field installable guides to increase their holding power. A word of caution is necessary, especially with the field-installable guides, to make sure the rods are slowly run in or out of the well to decide if a wellhead running guide is necessary.
Molded-on rod guides are the recommended type, especially for new sucker rods, if continued rod coupling/tubing wear is a problem. This type of guide is also recommended if the well is allowed to pound fluid or if the well-servicing contractor is not properly trained to run rods with field-installable guides.
There are varieties of materials that can be used for rod centralizers, including steel paraffin scrapers. However, most guides and scrapers are elastomers, including rubber, nylon, isobutyl, Ryton PPS (polyphenylene sulfide)*, a nylon composite, and a high-density polyethylene. Guide manufacturers continue to develop new guide materials that will provide the needed centralizing capabilities, rod-gripping strength, long wear life, and ability to function in increasingly hostile downhole environments. All these materials have chemical compatibility, temperature, and applied-stress limitations. The manufacturer should be consulted for their recommended service limitations.
Mechanical scrapers fastened to the rod string through the zone of paraffin deposition (normally near the surface) have been used to keep the tubing and most of the rod bodies free of paraffin. Paraffin-scraper systems have proved to be effective in reducing, if not eliminating, hot-oiling or watering treatments in both Canada and in the U.S. Additionally, a Canadian operator has shown that, along with the mechanical scraper system, internal plastic tubing coating has been beneficial in preventing paraffin buildup. However, it is recommended that paraffin scrapers be used only when necessary.
Sucker-Rod Pumping Units
Many devices are connected to the downhole sucker-rod equipment through the polished rod on the surface that imparts the reciprocating motion to the rod string and pump. In the history of sucker-rod pumping, a standalone, surface-pumping unit has become the proven technology. Many pumping-unit types are commercially available. Those most widely used have a walking beam as the horizontal load-bearing element and a sampson post that vertically supports the beam. These terminologies and configurations were adapted from the cable-tool drilling rigs used to drill early oil wells and developed into the conventional pumping unit.
API has standardized the design, terminology, and many components used for pumping units in API Spec. 11E. ISO accepted the use of this standard as a base to fast track the publication of ISO Standard 10431. Currently, these are comparable standards and cover the two main components making up a pumping unit: the gear reducer and the structure. They are standardized separately because the gear-reducer manufacturer may be separate from the structural manufacturer, who would be responsible for the assembly.
A pumping unit results when the gear reducer and the structure are combined together. These units have a size rating that describes the unit's capacities with the reducer rating, maximum structural capacity, and the maximum stroke length. The reducer number is the maximum torque rating in lbf-in. divided by 1,000. The structure number is the maximum load normally on the beam in lbf divided by 100, while the maximum stroke length is in inches. This results in a three-number hyphenated description that ranges from 6.4-21-24 to 3,648-470-300 for the 77 possible standardized units. These describe the smallest unit with a 6,400-lbf-in. reducer, a 2,100-lbf structure capacity, and 24-in. stroke to the largest unit with a 3,648,000-lbf-in. reducer, 47,000-lbf structure, and 300-in. stroke. However, not all of these unit sizes are available from all manufacturers in all the possible structural geometries.
The commercially available units are further described by adding the structural type or geometry and possibly the type of gear reducer [single (no letter) or double (D)]. Normally,
- B is for a beam-balanced conventional unit.
- C is for a conventional crank-balanced unit.
- A is for an air-balanced unit.
- M is for a Mark II unit.
- RM is for Reverse Mark * unit.
An example designation for a conventional, crank-balanced pumping unit with a 456,000-lbf-in. double-reduction-gear reducer, a 30,500-lbf structure, and a maximum stroke length of 168 in. would be C456D-305-168.
Manufacturers should be contacted for their normal availability, special designs, sizes, and types of units they sell. However, Table 11.7 shows the minimum and maximum size ranges commercially available from a large U.S. manufacturer.
Gear ReducerThere are 18 gear-reducer sizes currently included in API Spec. 11E. The size range is from 6.4- to 3,648- or 6,400- to 3,648,000-lbf-in. capacity. Table 11.8 presents the various sizes and capacities of available API gear reducers. When these gear reducers are put in their operating enclosure and attached to a pumping-unit structure, then this equipment is normally called a gearbox. Pumping units typically use single- or double-reduction gearing, with an approximate 30:1 speed reduction from the prime-mover to the pumping speed.
The standards also include chain reducers that use sprockets and chains for transmitting the prime-mover speed through the structure to the rod string. These are available as single-, double-, and triple-reduction drives. While this is still a possible reducer design, they are limited in capacity and are not normally used.
Gear Ratings for Speed and LifeSucker-rod pumping units can be operated over a range of pumping speeds. It has been recognized that there is a need for a nominal pumping speed to rate the various gear reducers. Originally, the industry adopted a nominal speed of 20 spm. This assumed that the up and down stroke of a unit forms one complete stroke cycle.
In 1981, API Spec. 11E was revised and reduced the rating speed for the 456- and larger-sized reducers, as shown in Table 11.9. The reduced speed setting was done because it was not practical to expect larger gearboxes to operate at 20 spm with longer stroke lengths and larger-sized structures. In actuality, industrial applications with these similar-sized reducers can be operated from 580 to 1,750 rpm. American Gear Manufacturer's Association (AGMA) Standard 422.03, which is the basis for API Spec. 11E, limits the speed of the reducer to either the pitch-line velocity of any stage to 5,000 ft/min and/or the speed of any shaft to less than 3,600 rpm.
It should be noted that none of the industry standards from API, ISO, or AGMA address a required reducer life; however, the operating rule of thumb is an expected 20 to 25 years of life. This assumes the gearbox is not overloaded or abused and is properly maintained. One pumping unit manufacturer has developed a graph (shown in Fig. 11.9) depicting the effect on gearbox life from overloading the gearbox capacity*. This shows that, while current API designed and manufactured reducers may be overloaded without catastrophic failure, depending on the amount of overload, the expected life should be reduced.
AGMA Standard 2001-C95 provides a way to calculate tooth stress that should provide satisfactory operation for a reasonable time. If the existing calculations are used and worked backwards to calculate the life of an acceptable design, then a reducer life of more than 4 × 108 cycles should be expected at the rated torque load. This would result in a life—assuming a constant 10-spm pumping-unit speed for every day of the year—of more than 76 years. However, this still assumes proper gear-reducer installation, operation, and maintenance.
The industry standards for pumping units have developed minimum requirements for the design and manufacture of the various structured components—the beams, shafting, hanger, brakes, horsehead, cranks, and bearings. The four main standard pumping-unit structural geometries covered by API Spec. 11E are as follows:
- Rear-mounted geometry, Class I lever systems with crank counterbalance.
- Front-mounted geometry, Class III lever systems with crank counterbalance.
- Front-mounted geometry, Class III lever systems with air counterbalance.
- Rear-mounted geometry, Class I lever systems with phased-crank counterbalance.
These standardized structures are more widely known by the respective designations: conventional, Mark II, air balanced, and Reverse Mark. There are variations of these geometries, such as for slant wells or as low profile for overhead irrigated fields. Additionally, there are special geometries or structures that are based on hydraulics, pneumatics, or belts. Because these structures are not covered by industry standards, it is recommended that these special units are designed properly, manufactured to industry quality standards, and installed and operated according to the manufacturer's recommendations.
There have been many publications about the advantages, disadvantages, and selection of the various standard geometries and the specialty pumping units, including the following:
- Theoretical development of torque factors and pumping unit "kinematics." 
- Description of geometries, applications, and efficiencies for standard units.
- Specialty hydraulic, strand, pneumatic, and long-stroke pumping units.
The following paragraph provides a brief summary and comparison of the four standard pumping units.
The conventional unit is probably the unit used most often. It is simple to install, has the widest range of sizes available, usually has lower operating costs than other units, needs no hoisting equipment or rigid supports for changing stroke length, and can run faster in wells in which free fall limits pumping speed. The maximum pumping speed for the conventional unit in an average well is estimated at 70% of the maximum free fall of rods in air. This compares with 63% for air-balanced units and 56% for Mark II units. The free-fall speed is defined for the conventional unit by the following formula:
The free-fall speed is reduced by 10 and 20% for the air-balanced and Mark II units, respectively. This means that in a well with average friction and a 100-in. polished-rod stroke, the rods will fall a maximum of 17.15 spm with a conventional unit, 15.43 spm with an air-balanced unit, and 13.72 spm for the Mark II. However, there should be no separation between the carrier bar of the unit and the polished-rod clamp during the downstroke. These speeds would be further reduced in wells with increased friction from composite-ring-type plungers, deviated holes, particulates sticking the downhole pump, and/or very viscous crude. Furthermore, the conventional unit's geometry allows either clockwise or counterclockwise rotation. This may be beneficial for gear teeth that are damaged in one direction from poor operation or maintenance and may enable rotating in the opposite direction. This would extend the life of the gearbox.
Air-balanced units use a leverage system different from conventional units. The use of compressed air instead of heavy, cast-iron counterweights allows more-accurate fingertip control of the counterbalance, which can be adjusted without stopping the unit. With no counterweights, the unit weighs much less than a comparably sized conventional unit. It also has a lighter substructure and a slightly lighter beam. Thus, there are several advantages to its compact size and light weight, especially for portable test units and for use on offshore platforms. It also uses more degrees of crank travel to complete the first one-half of the upstroke, which tends to decrease the peak load. This is a slight advantage if rod fatigue is a problem. However, there are increased maintenance problems or concerns, especially with leakage past the piston, which may make it difficult to maintain the proper air pressure. Additionally, the leakage also may cause an oil spray and resulting environmental consideration. Further, water condensation in the air system may cause damage if it is allowed to freeze, unless proper antifreeze is used.
The Mark II unit has an equalizer bearing between the Samson post and the well load. The equalizer bearing is located ahead or to the well side of the centerline of the slow-speed shaft. This is different from the air-balanced unit in which the equalizer bearing is directly over the slow-speed shaft. The equalizer bearing location results in an upstroke of approximately 195° and a downstroke of 165°. This makes a slower upstroke with 20% less acceleration, which results in reduced peak polished-rod load. The slower upstroke also allows more time for viscous fluids to fill the pump barrel and can increase the pump's volumetric efficiency, but this requires the unit to operate only in the counterclockwise rotation.
While comparably sized Mark II units are heavier and more expensive than conventional units, the claimed torque reductions may make it possible to use a Mark II unit one size smaller than required for a conventional unit. However, these units should not be used when high pumping speeds or undertravel-type dynamometer cards are anticipated and/or there are crooked or deviated wells. When an undertravel card or a card that showed neither undertravel nor overtravel is developed, the conventional or Reverse Mark unit has a better-suited permissible-load diagram.
The Reverse Mark unit is classified as a rear-mounted geometry, Class I lever system with phased-crank counterbalance. The phased cranks improve load-lifting capabilities; thus, like the Mark II, this unit may enable a one-size-smaller gear reducer than a conventional unit. However, this rule of thumb needs to be tempered by the actual pumping parameters and resulting dynamometer-card shape. Furthermore, the phase crank also makes this a unidirectional unit.
The other specialty units have their own advantages and disadvantages that may be considered if the standard units are not capable of meeting production-design requirements. Regardless of which unit is selected, a full-cycle economic consideration should be conducted to compare the costs for purchase, installation, maintenance, operation, repairs, failure frequency, and resale value. These parameters should all be considered, along with the capability of producing the required fluid volume from the required well depth, to decide which unit would be best for a particular well.
SizingThere have been a variety of methods for determining the required reducer size for a pumping unit, including the "approximate method," "engineering analysis," and kinematics. Today, most engineers/operators who select the pumping unit will rely on the output from a rod-string-design program that calculates the peak torque at the polished rod. These are based on the API RP 11L method and the extension to wave equations that allow geometries other than the conventional unit to be considered. Because these calculations provide peak torques at the polished rod, the torque has to be transmitted through the structure and its bearings to the gearbox. However, because these bearings are not 100% efficient, Gipson and Swaim developed curves for selecting the gearbox to account for these inefficiencies; Fig. 11.10 shows the loss of efficiency curves for both new and used units. Typically, this requires a gearbox approximately 10 or 20% larger in capacity than the peak torque calculated at the polished rod for new or used units, respectively. Once the design's peak-torque capacity is determined, then the closest available, but higher-rated, reducer should be selected. The beam should be selected on the basis of the calculated peak polished-rod load from the rod-string-design program. Finally, the unit stroke length should be selected on the basis of the required pump capacity with a 10 to 20% production cushion.
Specialty pumping units and the required reducer, structural capacity, and the desired stroke length should be discussed with the manufacturer to guarantee unit performance.
Installation, Operation, and Maintenance of Pump Units
Many publications have been issued on the installation, operation, maintenance, and lubrication of pumping units. These papers have been incorporated into API RP 11G1 to reflect the minimum recommended practices considered for installation, operation, and lubrication of the pumping unit. Additionally, manufacturers of the units may have their own documents and recommended procedures for installation, operation, and maintenance that should be followed.
Properly guarding a pumping unit is of critical importance. The industry standard, American National Standard Institute (ANSI)/API RP 11ER, should be followed when guarding the pumping unit, V-belts, sheaves, flywheels, cranks, counterweights, and moving parts on pumping units. Major pumping-unit manufacturers are also excellent sources of guidance on guarding and can usually supply guards that will meet specific regulatory requirements.
* Mark II and Reverse Mark are registered trademarks of Lufkin Industries Inc., Lufkin, Texas.
- Personal communication with C. Hunt, Lufkin Industries Inc., Lufkin, Texas (2002).
The prime mover (PM) rotates the gear-reducer gears through a V-belt drive. The two most common PMs are electric motors and internal-combustion (IC) engines. The decision concerning which to use depends on a variety of considerations, which includes the following:
- Availability of the power source (electricity or combustible fluid).
- HP required to pump the well.
- Efficiency of the system.
- Ability to control the PM to match the on/off potential operation of the pumping unit.
- Availability of field and/or service personnel capable of maintaining and repairing the equipment.
- Condition of the gas (sweet or sour) or availability now and in the future of the gas or liquids (i.e., propane or diesel) if an IC engine is used.
- Current and future expected cost for the power source.
- Anticipated full-cycle total cost (including initial capital, operating, maintenance, downtime, and repairs) for the duration of the well.
There are three common types of gas engines used for beam pumping units: two-cycle, slow-speed engine; four-cycle, slow-speed engine; and four-cycle, high-speed engine. The characteristics of these engines are summarized here, and the detailed comparisons and field experiences have been published elsewhere.
Two-cycle, slow-speed engine (less than 750 rpm):
- A minimum number of moving parts.
- Rugged, heavy-duty construction.
- A heavy flywheel that provides comparatively uniform crankshaft rotation on the cyclic loading of a pumping unit.
- Requires a minimum amount of maintenance.
- Can be overhauled on location.
- Requires a heavy foundation.
- Higher cost per HP than for high-speed engines.
- Weight per HP is higher than for high-speed engines.
- Can usually run only on natural gas or liquefied petroleum gas (LPG).
- May have either one or two cylinders.
- Fuel-injection system should be used when HP is greater than 40.
Four-cycle, slow-speed engine:
- Widely used.
- Relatively few moving parts.
- Uniform crankshaft speed because of a large flywheel.
- Can operate on governor control to compensate for load changes.
- Will operate on either natural gas or LPG.
- Repairs can usually be made without removing the engine from the pumping unit.
- Cost and weight per HP is greater than for high-speed engines.
- Limited engine sizes.
- Usually has a single horizontal cylinder.
Four-cycle, high-speed engines (greater than 750 rpm):
- Best suited for portable test installations vs. permanent installations.
- Lower initial cost.
- Lower weight per HP.
- Wide speed and power range.
- Operates on a variety of fuels.
- Large speed variations occur during pumping cycle because of a small flywheel effect.
- Operates on a fixed throttle with the governor mechanism acting only as an overspeed device.
- Has relatively short life because of the fast moving parts and the close tolerances required.
- Requires frequent oil changes.
- Requires frequent maintenance.
- Major repairs require that the engine be removed from the pumping unit.
API Spec. 7B-11C contains standard test and operating procedures that are used by manufacturers to determine the ratings of engines for oilfield service. These test data should be requested and furnished to the purchaser from the manufacturer. The data should include the manufacturer's curves showing the torque, maximum brake HP, and the rated-brake HP vs. engine speed. These are important to know the speed range in which the engine would be able to operate.
A general guide for installation and maintenance of gas engines is API RP 7C-11F, which covers all three types of engines and includes a troubleshooting section. This practice should be used as a starting point for engines unless the specific manufacturer's operating manual details otherwise. Additionally, there are a number of published papers on installation, care, operation, and lubrication of engines as prime movers for pumping units.
Gas-engine performance needs to be derated for altitude and temperature. The API Spec. 7B-11C for IC engines recommends the following:
- Deduct 3% of the standard brake HP for each 1,000-ft rise in altitude above sea level.
- Deduct 1% of the standard brake HP for each 10° rise in temperature greater than 60°F or add 1% for each drop in degree, if temperature is less than 60°F.
- Deduct 20% if the engine is continuously operated.
One of the biggest drawbacks of using IC engines is being able to automatically control their operation. There have been a few publications on automatic controllers, but these typically have had limited field use with no long-term production performance recorded.
Once it has been determined that an electric motor is needed vs. a gas engine, there are several things to consider, including design standard, unit efficiency, cyclic-load factor, and motor enclosure. These factors are discussed later in this chapter. Additionally, there have been a number of papers written on the use of electric motors for sucker-rod-lifted wells. Detailed discussions with example problems for sizing motors, along with discussion of electrical-power distribution systems for multiple-well installations, are presented in previous editions of the Petroleum Production Handbook and the Petroleum Engineering Handbook.
The electric motor most commonly used for beam-pumping installations is an alternating-current (AC), three-phase, squirrel-cage induction motor. These motors are used for the following reasons:
- Suitability for the load requirements.
- Low initial cost.
- Service dependability in the field.
If three-phase power is not available, single-phase motors up to 5 HP can be used. This motor is larger and more expensive than the three-phase motor of the same HP. The amount of motor voltage (V) needed depends on V on the distribution system, distance to the transformers, and motor size.
A general guide of motor size vs. V is 115 or 230 V for single-phase motors; 115, 230, 460, or 575 V for polyphase motors up to 50 HP; and 460, 575, or 796 V for polyphase motors 50 to 200 HP. Motors for pumping units come in a variety of common sizes: 1, 1.5, 2, 3, 5, 7.5, 10, 15, 20, 25, 30, 40, 50, 60, 75, 100, and 125 HP.
Natl. Electrical Manufacturers Assn. (NEMA) Design Standards
Motors can be purchased in six standard synchronous speeds, with the 1,200-rpm motor being the most commonly used in oilwell pumping. Multiple-HP-rated motors that may be either dual- or triple-rated are sometimes used for oilwell pumping; the triple-rated is more common. Changing one of these motors from one HP rating to another requires changing leads in the motor housing, which in turn changes the motor's internal wiring system. Any capacitors, fuses, or overload relays in the circuit will also require evaluation and possible revision at the same time to make sure it agrees with the new voltage/current requirements.
NEMA presents five general design standards that provide for varying combinations of starting current, starting torque, and slip. The most commonly recommended electric motor for pumping units is a 1,200-rpm NEMA Design D. It has a normal starting current, a high starting torque (272% or more of full-load torque), and a high slip (5 to 8%). Because Design D specifications are not drawn as closely as they are for other designs, manufacturers have developed several designs with variations in slip that still fall within Design D specifications.
The other NEMA designs (A, B, C, and F) are not used as often. However, there have been publications concerning when NEMA C and/or B designs could be considered, especially with variable-speed drives.
A power factor determines the amount of line current drawn by the motor. A high power factor is desirable because it is important in reducing line losses and minimizing power costs. A lower power factor means that the unit is not operating as efficiently as it should. Oversized motors tend to have low power factors. Typically, a NEMA D has a power factor of 0.87 when fully loaded, but decreases to 0.76 at half load. Usually, units must operate at a power factor of greater than 0.80 to avoid penalties from the power companies; thus, optimization of the pumping unit's size and motor needs to be considered as the well-fluid volume changes.
Using capacitors can increase power factors. To determine if and how much capacitance is needed, determine the power factor of an installation upon initial startup and then decide if a correction is justified. If a pumping-unit motor has a low power factor, a capacitor can be placed between the motor and disconnect. Because of the possibility of electrical shock, only qualified personnel should make this connection. Remember that changing producing conditions might require that the power factor be checked and that the motor-overload relays be resized if the capacitor is on the load side of the overload relays.
When a motor is used for a cyclic load, such as oilwell pumping, it will be thermally loaded more than the same average load applied on a steady-state basis. HP ratings of electrical motors depend on how much the temperature increases in the motor under load. A motor functioning cyclically must be derated from its full-load nameplate rating.
A motor's true performance and rating on a cyclic-load application cannot be determined by the use of normal indicating- or recording-type instruments. Motor heating is a function of the thermal current or root-mean-square (RMS) current, which is the square root of the mean of the squares of currents of definite time intervals. This may be more easily determined with an RMS or the thermal-type ammeter, which records RMS current corresponding to the true heating or "thermal" HP load on the motor. This current will always be higher than the average input current. The ratio of the average HP output to the "thermal HP output" corresponding to the RMS line current is called the motor derating factor and is always less than one. Its inverse is the cyclic-load factor, which is always greater than one. An average motor derating factor for NEMA Design C motors is 0.65; an average motor derating factor for NEMA Design D motors is 0.75.
There are four basic types of motor enclosures: drip-proof guarded, splashproof guarded, totally enclosed fan cooled (TEFC), and explosion proof. "Guarded" refers to screens used over air intakes to prevent the entrance of rodents or other foreign items. The TEFC enclosure provides the maximum protection for the interior of the motor. The drip-proof motor should prove adequate for most pumping-unit installations in which the motor is elevated. This type of construction is built with a closed front-end bell to eliminate the entry of horizontal rain, sleet, or snow into the motor. The splashproof motor affords somewhat more protection against splashing liquids than does the drip-proof one. The preferred enclosure sets the motor on or close to the base; the explosion-proof enclosure will seldom be required. Motor-high mounts on pumping units have also been useful in protecting the motor from sand or snow.
NEMA has established the insulation classes and the maximum total temperatures applicable to these classes for insulations used in motor winding. For normal service life, the temperature of the motor windings should not exceed the maximum allowable temperature for that particular insulation type. Class A insulation has a maximum total temperature of 105°C, Class B = 130°C, Class F = 155°C, and Class H = 185°C. Generally, the more the motor enclosure restricts the flow of outside cooling air, the higher the temperature rise will be, and in all probability, the higher the winding temperature. This temperature increase has to be incorporated into the decision regarding which insulation class is required.
The service life of an AC induction motor is determined by the bearing life, the insulation life, and routine maintenance/inspection. Temperature rise is important because studies have indicated that for every 8°C rise above the temperature values stated, the insulation life is cut approximately in half.
Slip is the difference between motor synchronous speed and speed under load, usually expressed in percent of synchronous speed. Synchronous speed is the theoretical, no-load speed of the motor. Slip characteristics are very important because they will determine how much HP can be converted to torque to start the gearbox gears turning. A high-slip motor permits the kinetic energy of the system to assist in carrying the peak-torque demands. A low-slip motor will respond to the instantaneous demand; in other words, the high-slip motor slows down more under peak torque demands than the low-slip motor. The result is that the high-slip motor will require lower peak currents than the low-slip motor. How high the motor slip should be for pumping installations is debatable; however, Howell and Hogwood stated, "A slip greater than 7 to 8% offers no additional advantages from the overall pumping efficiency standpoint."  On the basis of this information and the slip characteristics of the various designs, the Design D motor with a 5 to 8% slip is recommended for most sucker-rod installations.
Ultrahigh-Slip (UHS) Motors
Higher-slip motors are available from some manufacturers; one has claimed to have slip characteristics up to 35 to 40%, also claiming that using their UHS motor would result in lower loading on the sucker rods, lower electric-current peaks, and reduced power use. However, to obtain the mechanical advantage, these systems have to be set up in the high-slip mode. When this is done, the increased slip normally decreases the operating speed and may result in a decrease in production when compared to a NEMA D installation.
Motor controls are housed in a weatherproof, NEMA Type 3 enclosure with special explosion-proof enclosures available. All control units should contain the following:
- Fused manual disconnect.
- Hand on/off/automatic selection switch.
- Lightning arrester system.
Circuit breakers are sometimes used instead of fuses. The fused manual disconnect acts as a line-disconnect switch at the entrance to the control box. A fused disconnect may be located on a pole upstream of the motor starter; the lightning arrester is connected to the incoming line terminals, just ahead of the fused-manual disconnect and must be properly grounded. Depending on the inherent protection built into the motor, the control box may contain an overload relay, an undervoltage relay, and/or a sequence-restart timer.
The electrical equipment must be properly grounded. Good grounding procedures are essential to personnel safety and good equipment operation. It is recommended that reference be made to the Natl. Electrical Code and the Natl. Electrical Safety Code to ensure safe grounding is met. Particular attention should be given to the connection of the ground wire to the well casing. The connection should be located where it will not be disturbed during well-servicing operations and should be mechanically secure. Periodic (yearly is recommended as a minimum) continuity measurements should be made with a volt-/ohmmeter between "a new clean spot" (not where the ground wire is terminated) on the well casing and new spot on each piece of grounded equipment. The resistance measured between any piece of equipment and the casing should not exceed 1 ohm. The resistance measured between the pumping-unit ground system and another nearby moisture ground should not exceed 5 Ω. However, these measurements should to be checked with current circulating through the system to determine if the ground is good.
There are seven HP values that should be considered in the proper design and operation of sucker-rod-pumped wells; these are hydraulic, friction, polished-rod, gear-reducer, V-belt drive, brake, and indicated.
Hydraulic HP (HHP) is the theoretical amount of work or power required to lift a quantity of fluid from a specified depth. This is a theoretical power requirement because it is assumed that there is no pump slippage and no gas breakout. The HHP, thus, is the minimum work expected to lift the fluid to the surface and can be found with the following equations:
Friction HP (FHP) is the amount of work required to overcome the rubbing-contact forces developed when trying to lift the fluid to the surface. This friction can be caused by a number of sources including plunger-on-barrel friction; rod- and/or coupling-on-tubing wear; sand, scale, and/or corrosion products hindering pump action, rods, and couplings moving through the fluid; fluid moving up the tubing; normal and excessive stuffing-box friction; and liquid and gas flowing through the flowline and battery facilities. FHP, thus, is dependent on factors such as how straight and deep the well is, the fluid viscosity, the pumping speed, and the tubing/rod buckling. In most situations, unless we know all of these factors, we do not know what FHP is. However, for design purposes, API RP11L calculations assume the friction effects, which show up in the peak and minimum polished-rod loads and in the calculation of polished-rod HP (PHP).
PHP is the amount of work required to artificially lift the fluid to the stock tank. It is the sum of HHP plus FHP. For design purposes, API RP11L assumes these values are related to Fo/SKr and N/No, where Kr is the load necessary to stretch the rod string 1 in., and No is the natural frequency of a straight rod string. If a surface dynamometer card is available, the PHP can be measured because the area of the card is the work done at the polished rod to lift the fluid to the surface. The formula for calculating PHP follows:
Gear-reducer HP (GHP) is a value used to find the efficiency of the unit (i.e., how much the gear reducer is loaded, compared to required peak torque). GHP can be calculated by the following:
V-belt-drive HP (VHP) is the maximum power required by the V-belts to be transmitted to the gear reducer. API Spec. 1B states that the VHP for a beam-pumping unit is as follows:
Brake HP (BHP) is the power required by the prime mover to turn the sheave that makes the reducer's gears turn and starts the cranks going around. This power must accommodate the inefficiencies of all components involved in getting the cranks to turn to transmit the power to the polished rod. BHP can be found with Gipson and Swaim recommendations by the following equation:
The efficiency factor is found from a graph by taking GHP divided by API gearbox-torque rating and then intersecting either a worn- or new-unit efficiency curve. This efficiency factor is applied to the PHP to convert it to BHP at the prime mover and is required to offset power losses caused by friction in the surface equipment. Fig. 11.10 is a recommended curve to find the HP efficiency factor.
Additionally, a minimum estimate for this HP by NEMA for Design D and C motors is as follows:
This derating factor is 56,000 or 45,000 for D or C motors, respectively.
Indicated HP (IHP) is the power required by the prime mover to meet the BHP requirements and determines the size of motor that needs to be ordered. It is found through the following equation:
This derating factor accommodates continuous operation and thermal effects. The derating factors for electric motors are 0.75 and 0.65 for NEMA D and C, respectively. The derating factor for a gas engine is dependent on the type of engine and service, rotational speed, elevation, and ambient temperature. The effects of these parameters are discussed in API Spec.7B-11C, paragraphs 2.11 and 2.13. A rule-of-thumb estimate for an engine's derating factor is as follows:
HP Problem-Solving Example
Given the previous HP definitions, along with the information and calculations in API RP11L (p.7), find all seven HPs:
- HHP = [175 (BFPD) × 350 (lbf/bbl) × 0.9 × 4,500 (ft)] / (33,000 × 1,440) = 5.2 HP.
- PHP = line 26 = 8.5 HP.
- FHP = PHP – HHP = 8.5–5.2 = 3.3 HP.
- GHP = line 25/4,960 = 133,793/4,960 = 26.9 HP.
- Assuming a 160,000-lbf-in. unit is ordered to accommodate a calculated 133,793-lbf-in. peak torque, and using Fig. 11.10 , find the efficiency factor of 0.86: VHP = (133,793 × 16) / 70,000 = 35.6 HP.
- BHP = (PHP / efficiency factor), where the efficiency factor is found by GHP / reducer rating = (8.5 × 4,960) / 160,000 = 0.2635. With Fig. 11.10 , the efficiency factor is 0.64. Thus, BHP = (8.5 / 0.64) = 13.28 HP.
- Assuming a NEMA D motor, IHP = (BHP / derating factor) = 13.28/0.75 = 17.7 HP.
Therefore, a 20-HP motor should be purchased. However, a 15-HP motor may work, but certain aspects are not known, including actual counterbalance divided by optimum counterbalance, flowline pressure, and actual friction effects. Thermal current (amps) can be measured to determine how much motor capacity is actually being used once the unit and motor are installed. The actual motor size could then be refined for other units in the area.
Sheaves and V-Belt Drives
Prime movers—whether with a gas engine or an electric motor—run at a speed of 300 to 1,200 rpm. This speed must be reduced to the required pumping-unit speed of 2 to 25 spm. This is accomplished with sheaves, V-belt drives, and gear reducers. A sheave is a grooved pulley, and its primary purpose is to change the speed between the prime mover and the gearbox. The belt—usually a V-belt —is a flexible band connecting and passing around each of the two sheaves. Its purpose is to transmit power from the sheave on the prime mover to the sheave on the pumping unit. It is important to understand the basics of sheaves and V-belt to know how to select a sheave for a certain pumping speed and to determine the number of V-belt needed.
Sheaves come in different widths and have from 1 to 12 grooves. They are selected on the basis of the pitch diameter (PD) relative to how many spm the unit will pump. New beam-pumping units can be purchased with different-sized sheaves on the reducer. Sheaves can also be purchased to accept different V-belt cross sections. A pumping-unit sheave should be selected that will allow as much speed variation (up and down) from the design speed as is practical without violating API Spec. 1B rules. Most unit sheaves will have grooves for more belts than are actually needed because most units seldom, if ever, operate at maximum HP. The maximum VHP is shown in Eq. 11.15. Only the grooves closest to the prime mover and the gear reducer should be filled, and only enough belts to transmit the VHP should be installed because of the following considerations:
- The tension in the excessive belts, which will be further from the equipment than the required belts, will place unnecessary loads on the bearings.
- Wider sheaves than necessary and extra belts increase investment costs.
- It takes more energy to flex the extra belts around the sheaves, which increases operating costs.
Pumping-unit manufacturers usually list all unit-sheave sizes in their catalogs. Motor sheaves are available with various PDs and numbers of belt grooves. Table A.1 in API Spec. 1B contains commonly available sheaves. Because of availability, motor sheaves should be selected from those listed in the top portion of the table.
V-Belt BasicsA V-belt has a trapezoidal cross section that is made to run in sheaves with grooves that have a corresponding shape. It is the workhorse of the industry, available from virtually every V-belt distributor, and it is adaptable to practically any drive. It was designed to wedge in the pulley, thereby multiplying the frictional force produced by the tension; this, in turn, reduces the belt tension required for an equivalent torque. Remember, the purpose of the belt is to transmit power from the sheave on the prime mover to the sheave on the pumping unit. Therefore, the number and size of the belts needed depend on the amount of power to be transmitted.
Reinforcing cords normally made of rayon, nylon, or other polymer materials provide the load-carrying capability of a V-belt. The cords are usually embedded in a soft rubber matrix called a cushion section. The balance of the belt is made of harder rubber, and the entire section is usually enclosed (i.e., wrapped) in an abrasion-resistant jacket or cover.
As the belt bends around a sheave, the bending-neutral axis is the only portion that does not change the circumferential length. This line (which does not change length) is called the pitch line and determines the "effective" radius of the pulley, which in turn, determines the torque and speed ratios. The position of this line as it curves around the pulley forms a pitch circle with a pitch diameter.
Classical V-belts are made in five standard cross sections designated by the letters A (the smallest cross section), B, C, D, and E (the largest cross section). The HP that a belt is able to transmit falls off rapidly as the sheave size diminishes. Table 11.10 lists the minimum PDs recommended by API for the various belt sections. Smaller-PD sheaves are not recommended because of decreased HP, reduced transfer efficiency, shorter belt life, and less economical drive. Fig. 11.11 shows the HP capacity a single belt can transmit for a selected small-diameter sheave for the various belt cross sections.
Other Types of Belts
There are other types of belts (i.e., flat, narrow, and synchronous belts, as well as other variations of the V-belt). For example, narrow multi-V-belts (power bands) were developed because the maximum load capacity for a given width of belt required the use of a narrow section. This provided the maximum support of the tensile cords by joining the belts together. V-ribbed belts provide complete support with only a modest compromise in terms of additional tension.
Selecting a Sheave
The first step in designing the V-belt drive for a pumping unit consists of selecting a sheave for the unit and the prime mover. To do this, the desired pumping speed (N), along with the speed (in rpm) of the prime mover and gear ratio, must be known. If the other parameters are known, this equation can be rearranged to determine any required factor:
The largest motor sheave in this group will provide for the greatest reduction in pumping speed for future operations merely by changing motor sheaves.
Double Reduction With Electric Motor
A double-reduction unit run by an electric motor will require a speed reduction through the V-belt drive of approximately 2:1 at fast pumping speeds. At slow speeds, the ratio will be 6:1. When two belt sections are offered for the unit sheave, the smaller belt section will allow the use of a smaller motor sheave and a lower pumping speed. In most cases, the smaller belt section, with one of the two largest-unit sheaves, will offer the greatest flexibility.
Double Reduction With Gas Engine
A double-reduction unit run by a slow-speed gas engine will require a speed reduction of 1:1 at a fast pumping speed; at a slow pumping speed, the ratio will be 3:1. In these cases, speed reductions (which may be anticipated through the drive) should be checked with the proposed unit and prime mover. If little or no speed reduction will ever be required through the V-belt drive, one of the two smaller-unit sheaves will enable the use of a smaller (and less-expensive) prime-mover sheave. The larger belt section could also be used and may require fewer belts.
Determining the Required Number of Belts
The first step in determining the number of belts required is to calculate the VHP. When the peak torque is known, this is the preferred method of calculating the design HP. When the peak torque is not known, a service correction of 1.6 is recommended.
The remainder of the calculation can be performed by following the procedure in Section 4 of API Spec. 1B, starting with paragraph 4.5 (page 11). A complete design requires that the distance between the centers of the driver and driven sheaves be known. The basic steps are given in API Spec. 1B. An example calculation is presented here.
As an example problem, select the optimum gear-reducer sheave for a C-160D-173-86 pumping unit that will be operated with the reducer fully loaded.
Given: gear-reducer sheaves available from the pumping-unit manufacturer's catalog: 20-, 24-, 30-, 36-, and 38-in. PD-3C. Assume that the prime mover's average rpm = 1,120. The smallest C-section motor sheave that should be considered = 9 in. PD (i.e., 9.4-in. OD in Table 3.1 of API Spec. 1B). The largest sheave that should be considered to keep the design PD velocity at less than 5,000 ft/min = 16-in. PD (calculations indicate a 17-in. PD, but page 32 of API Spec. 1B indicates that 17-in. PD C-section sheaves are not generally available; economics should discourage engineers and others from recommending sheaves not listed). The liquid to be pumped has a viscosity of approximately 1 cp. The pumping-unit gear ratio is 28.67. The maximum speed with an 86-in. stroke should result in an acceleration factor of 0.3, in which the maximum spm ≤ (0.3 × 70,500/86) 0.5 ≤ 15.7. The minimum speed with an 86-in. stroke should result in an acceleration factor ≤ 0.225, in which the minimum spm ≤ (0.225 × 70,500/86) 0.5 ≤ 13.6.
Find: the optimum gear-reducer sheave and the number of C-section belts required, assuming the reducer is fully loaded and is operated at the maximum and minimum speed dictated by the sheave selected.
The table shows that the 38-in. PD-4C gear-reducer sheave should be selected; however, the 36-in. gearbox sheave is acceptable.
Solution 2: 1. VHP at 9 spm = 160,000 × 9/70,000 = 20.6. 2. HP that can be transmitted with one C-section belt and with a 9-in.-PD prime-mover sheave (as shown in Fig. 11.11) = 11. 3. Number of belts required = 20.6/11 = 2 belts. 4. VHP at 16 spm = 160,000 × 16/70,000 = 36.6 5. HP that can be transmitted with one C-section belt and with a 16-in.-PD prime-mover sheave (as shown in Fig. 11.11 ) = 25. 6. Number of belts required = 36.6/25 = 2 belts.
Note that neither calculation justifies filling all the grooves in the gear-reducer sheave. No justification is known for using more belts than is indicated by API Spec. 1B.
Miscellaneous Surface Equipment
A polished rod is the top-most rod in a rod string. These rods come in various lengths and sizes. Polished rods are made of various materials, including carbon steel, stainless steel, and monel. It is usually more economical to use corrosion-resistant polished-rod liners on carbon-steel polished rods than to use corrosion-resistant polished rods. Polished rods must be properly aligned in relation to the pumping tee. Poor alignment will result in decreased life of the stuffing-box packing and possible failure of the polished rod. Furthermore, if the polished rod does not travel straight up and down during the pumping cycle, liners may not be practical. For situations in which the pumping unit is not properly set and/or the wellhead is crooked, a full-length sucker rod should be installed between the polished rod and the top of the string's pony rods. This will decrease crooked wellhead-induced polished-rod failures and increase packing life. The polished rod must have a coupling and a sub on top. This is required in case the rod slips because the polished-rod clamp is not sufficiently tight. The coupling keeps it from falling through the stuffing box. The subrod helps retrieve the polished rod and helps prevent moisture from getting into the coupling.
Section 12 of API Spec. 11B discusses polished rods and polished-rod liners. Table 12.1 in API Spec. 11B recommends polished-rod size vs. the size of the top rod in the rod string. API polished-rod lengths are 8, 11, 16, and 22 ft. Upset ends can be furnished on 1 1/8-, 1¼-, and 1½-in. polished rods and are recommended for heavy loads. Upset ends have sucker-rod connections that are superior to the pipe-thread connections on nonupset polished rods. This type of connection decreases stress concentration and results in improved fatigue life. The surface finish on polished rods is specified in Section 12 API Spec. 11B. Although the range of surface finish is 10 to 20 micro-inches, roughness average scale (RA), it is recommended that a 16-micro-inch- RA finish be specified because, if the finish is too smooth, it may be difficult for the clamps to work properly and a too-rough finish reduces polished-rod packing life.
Polished-rod clamps are fitted on the polished rod and come in several designs. Clamps for the light loads may have only one bolt, whereas clamps for heavier loads will have two bolts. The clamp manufacturer specifies the torque required to tighten the clamps, which is also discussed in both API Spec. 11B and API RP 11BR. They also specify the forces that will cause clamps to slip on polished rods in API Spec. 11B. This is based on the assumption that the OD of the polished rod will be approximately equal to the OD the manufacturers assumed when they designed and built the clamp. The clamp must be the right size for the polished rod (no homemade bushings) and be strong enough to support the maximum well load. Open-end, box-end, or socket wrenches should be used on the clamp nuts and bolts. Pipe wrenches cut the nuts and make it hazardous for those who must loosen the clamp in the future. Be careful of foreign material in the clamp or on the polished rod. If the polished rod and clamp are not properly cleaned, the clamp may slip. Clamps that do not have a load-bearing surface perpendicular to the polished rod can also bend the polished rod. The following are some maintenance tips to keep in mind when working with the clamps:
- Use the clamp manufacturer's recommended torque for tightening the bolts. Do not overtighten polished-rod clamps—it may be the start of polished-rod failure. API Spec. 11B requires that a properly attached clamp may not cause an indentation of more than 0.010 in.
- The polished rod's clamp area and the inside area of the clamp should be cleaned before installation.
- Do not allow the use of pipe wrenches on polished-rod bolt nuts. Replace all pipe-wrench-cut nuts.
- Do not put clamps on polished-rod liners.
- Do not clamp on the sprayed-metal part of polished rods.
A stuffing box is a device attached to the pumping tee that seals fluids in the tubing by forming a tight seal with the polished rod and diverting the produced fluids out of the pumping tee into the flowline. Packing for stuffing boxes is made from a variety of different materials. Local experience is the best guide in selecting the appropriate packing material to use.
Stuffing boxes may have one or two sets of packing elements. In a stuffing box with two sets of packing, the lower set is left relaxed and inoperative during normal operations. When it becomes necessary to replace the upper set of packing, the unit is shut down, and the lower set of packing is tightened against the rod, which enables the upper-packing element to be safely replaced with pressure on the tubing. After replacing the upper element, the lower-packing element must be backed off before starting the unit. This method not only retains the tubing pressure and decreases pollution, but also keeps low-pressure gas out of the face of the person doing the work.
There are stuffing boxes made with attached oil containers to keep the polished rod lubricated on wells that pump off, have high water cuts, or are in a semiflowing gas-heading condition. The proper method for handling the pumpoff condition is adjusting the pump capacity with time clocks, stroke lengths, stroke, speed, or pumpoff controllers. Maintaining a surface backpressure on the tubing may be beneficial on wells that are in a semiflowing gas-heading condition. Both conditions should be corrected to decrease polished-rod and stuffing-box wear and to increase overall pumping efficiency.
Rod rotators must be used with certain types of mechanical paraffin scrapers. Rod rotators may also be used when rod-coupling wear is a problem. The rotation of the rods spreads the wear around the entire surface of the coupling instead of allowing it to be concentrated on one small area. Rotation does not solve the problem, but it does make the coupling or centralizer last longer. Rotators need to be selected properly and are dependent on the well load.
API Spec. 11B covers design and rating of pumping tees. The major requirement for tees and stuffing boxes are that they be properly installed. In addition, the threads need to be clean and in line with the tubing when it is screwed on.
A check valve is a valve that permits flow in only one direction. If the gas or liquid flow starts to reverse, the valve automatically closes and prevents reverse flow. A check valve should be placed between the casing head and flowline to prevent backflow from the flowline into the casing annulus. An oversized check valve will chatter and destroy the seat seal prematurely; an undersized check valve will hold too much backpressure on the casing.
The casing/tubing annulus should be equipped with a wing valve that will allow the casing pressure and the fluid level to be monitored. This valve also can be used to introduce to the well corrosion inhibitors, hot oil, water, etc. It should be bull-plugged closed when not in use. Introducing liquids into the annulus at a higher rate than the annulus self-venting rate drives the producing-liquid level to less than the pump intake, which starves the pump and causes premature pump failure. Self-venting can occur if the equivalent annulus diameter ≥ 0.92 × Q0.4, where Q is the pumping rate in gal/min. Wing valves allow the installation of a pressure gauge so that casing pressure can be measured. This is important to check because, if the casing pressure is greater than ½ the pump-intake pressure, the flowline is probably too small or partially blocked.
Another type of surface valve that could be used is a backpressure valve. This valve is normally installed in the flowline, upstream from the casing-annulus gas-piping tie in and is typically used to keep the tubing from unloading when the well still has high bottomhole pressure (when the well alternates between flowing and pumping, this situation is called "flumping"). The optimum backpressure to prevent flumping would be equal to or just greater than the pump-intake pressure. It should be noted that backpressure on the tubing can cause paraffin deposits in the tubing to come loose, flow up the tubing, and block the backpressure valve, or may cause the stuffing-box packing to blow out; thus, the tubing and rods should be cleaned before applying backpressure.
There has been a long history of work trying to model or design sucker-rod strings. This includes the original work from Slonneger and Mills on vibration effects of rod strings. Fatigue of rods also was considered in 1940. These effects helped develop the Slonneger, Mills, and Langer formulas for rod loads. A detailed discussion and development of these formulas is provided by Zaba.
Zaba detailed the next refinement of sucker-rod loading, which was the organization of the Sucker Rod Pumping Research Inc. in 1954, and the development of an analog computer model to simulate the elastic behavior of rod strings. This method was provided to the industry in the 1960s, and the design results were developed into the hand-calculation and graphical method in API RP 11L.
Companies used this graphical chart and calculation method for many years, with some refinements and changes to the practice, to account for tapered-rod strings and rod percentages, that provide equal loading in each section of a string. The development of the wave equation for sucker-rod lift by S.G. Gibbs in 1961 was a major step forward because its use permitted design or analysis for all types of units and rod strings. The advent of the personal computer and its continued developments of power and speed allowed more developments of rod-string simulators, including extending the API simulator using a next-order wave equation, pumping units different than conventional ones, mixed-steel and fiberglass-rod strings, frictional effects of the fluid and wellbore deviation, and current models that address very viscous fluids and 3D horizontal wells. Regardless of what method or program is used to predict loads, once the equipment is installed and the well has stable production and fluid levels, it is recommended that a dynamometer survey be run with a load-capable dynamometer attached to the polished rod. The predicted loads should be compared to the actual loads and the associated fluid production. Adjustments to the predictions should be made for future troubleshooting and any further design changes.
While these models have improved, they still address only the loads on the selected grade of rods and the string design, the size of downhole pump, and the type and size of the pumping unit. However, for a complete design of a beam-pump installation, all the equipment discussed in the preceding sections needs to be addressed, as well as the data provided from a rod-string design program, which at minimum, include the following:
- Where the pump is set and the associated downhole separator design.
- Type of pump, along with its design and metallurgy.
- Sinker-bar use and design, if required.
- Tubing size and grade.
- TAC use, position, and setting.
- Polished-rod size.
- Polished-rod clamp size.
- Type and size of prime mover.
- Sheave and V-belt design.
Gipson and Swaim did an excellent job of summarizing a sucker-rod lift-system design in The Beam Pump Design Chain with the API RP 11L approach. This recommended practice should be consulted for continued discussion of this equipment, along with a review of a sample problem and a recommended solution. In summary, use the design procedure presented in API RP 11L or a suitable wave equation. Several commercial wave-equation computer programs are available that many operators have successfully used.
Automation and Pumping Control
Conoco unpublished internal report.
"Automation" means different things to different people and becomes a problem when the term triggers concern from the field about personnel reduction. Thus, sucker-rod-lift automation may not always be considered good if not properly applied. However, there needs to be monitoring and control equipment on an installation to enhance proper operation, monitoring, failure reduction or prevention, and troubleshooting/problem solving.
At minimum, a sucker-rod-lift installation should have vibration switches on the unit to shut it down if there is a high part in the rod string that will cause overloading of the gearbox or damage to the unit foundation. There should be a pressure gauge (or a connection for a pressure gauge to allow temporary installation) on the flowline-pumping T, downstream of the check valve that monitors the flowline pressure. There should also be some type of pump-cycle controller. This may be from a simple time clock to a more sophisticated pumpoff or rod-pump controller.
A number of papers have been published that address automation of sucker-rod-lift or beam-pump automation and control. There is also a reference on practical automation for mature fields. If a high degree of automation is considered, then a very important side consideration is keeping this electrical equipment working, especially during electrical storms; thus, proper lightning protection and grounding should be considered.
A study made several years ago indicated that at least one-half of the pumping wells surveyed had a subsurface pump installed that was too large.* The results of such installations were devastating fluid pounds when wells were overpumped, resulting in short run times and increased failure frequency. Because of the cost to pull and replace a pump, typically other parts of the sucker-rod-lift system were changed to compensate for the oversized pump. Too many times, the too-large pump is a result of habit or of not optimizing when the well capacity has changed.
It is still possible to live with the too-large pump until the correct size can be installed. Some interim measures are to reduce the pump displacement by reducing the strokes per minute, shortening the stroke, and decreasing backpressure on the tubing/casing annulus, thereby decreasing formation backpressure, allowing more fluid inflow, and reducing the pumping time.
Probably the most common type of well control or automation is time clocking, which consists of pumping a portion of a 15-minute period. Percentage timers and pumpoff controls are used in modern time-clocking work. The purpose of time clocking is to adjust the pump capacity to the well capacity.
Pumpoff controllers have been developed over the years to be standalone monitors, to provide rod-string load and polished-rod position and related dynamometer cards, and to be installed with communication links to allow remote monitoring and control of the installation. Current advancements in computers along with electrical end devices allow sophisticated control of individual installations and/or a whole field. If new pumping installations are planned, these types of controllers/automation should be considered. It becomes more difficult to justify a retrofit to a long-time producing field, but this may be considered depending on access to the field, variable well inflow, and/or reduction in operating costs by reducing well failures. Many papers on pumpoff or rod-pump controllers, different theories concerning their operation, and controller installation and operation have been published. These should be reviewed to determine if or when a controller may be advantageous to install.
Troubleshooting Sucker-Rod-Lift Installations
Once a sucker-rod-lift system is installed on a well, the continued monitoring and optimization of pumping parameters begins. Obtaining monthly well tests on the fluid production from the well and a fluid/pump submergence level is recommended to ensure that the well capacity is within the recommended pump-capacity range, the well does not have excess capacity or equipment needs to be changed because of excessively high fluid levels, and that excessive pumping of the well is not occurring.
Although current rod-string-design models, simulators, and programs are fairly accurate, they still need individual-well calibration to ensure that the design assumptions are correct for the actual well conditions. Additionally, to know what is different and why, the six main well loads need to be recorded from the predictive design. These loads need to be compared to the actual well loads with known fluid-level, well-test, and pumping parameters. Gipson and Swaim have described these six basic loads and their relationship to a surface dynamometer card.
Many papers have been published on dynamometers and their use on sucker-rod-lifted wells. Some of these provide discussion of surface loads and surface dynamometer cards, while the latest trend is to discuss downhole dynamometer cards (or pump cards). While obtaining actual downhole loads that these dynagraphs recorded, there has been recent work on developing and field-testing a downhole dynamometer.
While these measurements investigate the sucker-rod-string loads, the other components of the lift system also should be investigated, including the pumping unit and gearbox. As previously discussed, there are only two techniques to check if a pumping unit is overloaded: conducting a torque analysis or comparing the permissible-load diagram (PLD) for the pumping unit to the loads from the surface dynamometer card. The torque-analysis technique has been demonstrated by Gipson and Swaim, and Takacs, Gault, and Teel have discussed PLDs or envelopes. Chastain discussed examples of PLD use for properly counterbalancing a pumping unit.
Failures of sucker-rod-lift components have been discussed in countless papers. The use of current data processing and root-cause analysis of these failures has been the recent industry trend to assist in reducing failures., Additionally, the Artificial Lift Energy Optimization Consortium (ALEOC) program in west Texas has been useful for operators to compare the failure frequency of their sucker-rod-lift components, wells, and fields with other operators to find areas of improvement. One final new trend developing for this lift method is a total well-management concept that integrates the well capacity/pump submergence and rod-string and pumping-unit loads with power demands. This may prove the best practice for optimizing, troubleshooting, and reducing failures along with reducing associated lifting costs.
|a||=||casing/tubing annulus area, in.2|
|C||=||diametrical clearance between plunger and barrel, in.|
|D||=||plunger diameter, in.|
|Er||=||elastic constant rods, in./lbf|
|F||=||gradient correction factor|
|Fo||=||differential fluid load on the full pump-plunger cross-sectional area, lbf|
|Fo /SKr||=||dimensionless sucker-rod stretch load (fluid load on full plunger area divided by load necessary to stretch the total-rod string to an amount equal to the polished-rod stroke length)|
|G||=||specific gravity of the combined fluid in the tubing|
|H||=||pump seating depth, ft|
|VHP||=||V-belt drive horsepower|
|Kr||=||the load necessary to stretch the rod string 1 in.|
|L||=||pump-seating nipple depth, ft|
|Lp||=||plunger length, in.|
|LPSD||=||seating nipple/pump depth, ft|
|N||=||pumping-unit speed, spm|
|No||=||the natural frequency of a straight rod string, spm|
|p||=||differential pressure across plunger, psi|
|PD||=||pump displacement, BLPD|
|P||=||producing pressure, psia|
|Q||=||slippage or leakage loss, in.3/min|
|Q/aP0.4||=||parameter from Gilbert used to determine gradient correction factor, where Q is gas flow rate, Mscf/D; a is the casing-tubing cross-sectional area, in.2; and p is the producing pressure, psia|
|S||=||surface stroke length, in.|
|Sp||=||downhole pump-plunger stroke length, in.|
|WC||=||well production capacity, BFPD|
|μ||=||absolute viscosity of fluid, cp|