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PEH:Hydraulic Pumping in Oil Wells
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 14 – Hydraulic Pumping in Oil Wells
Hydraulic pumping is a proven artificial-lift method that has been used since the early 1930s. It offers several different systems for handling a variety of well conditions. Successful applications have included setting depths ranging from 500 to 19,000 ft and production rates varying from less than 100 to 20,000 B/D. Surface packages are available using multiplex pumps ranging from 15 to 625 hp. The systems are flexible because the downhole-pumping rate can be regulated over a wide range with fluid controls on the surface. Chemicals to control corrosion, paraffin, and emulsions can be injected downhole with the power fluid, while fresh water can also be injected to dissolve salt deposits. When pumping heavy crudes, the power fluid can serve as an effective diluent to reduce the viscosity of the produced fluids. The power fluid also can be heated for handling heavy or low-pour-point crudes. Hydraulic pumping systems are suitable for wells with deviated or crooked holes that can cause problems for other types of artificial lift. The surface facilities can have a low profile and may be clustered into a central battery to service numerous wells. This can be advantageous in urban sites, offshore locations, areas requiring watering systems (sprinkle systems), and environmentally sensitive areas.
Hydraulic pumping systems transmit power downhole by means of pressurized power fluid that flows in wellbore tubulars. Hydraulic transmission of power downhole can be accomplished with reasonably good efficiency using a reciprocating piston pump. With 30°API oil at 2,500 psi in 2 7/8-in. tubing, 100 surface hydraulic horsepower can be transmitted to a depth of 8,000 ft with a flow rate of 2,350 B/D and a frictional pressure drop of less than 200 psi. Even higher efficiencies can be achieved with water as the hydraulic medium because of its lower viscosity.
The downhole pump acts a transformer to convert the energy into pressure in the produced fluids. A common form of a hydraulic downhole pump consists of a set of coupled reciprocating pistons, one driven by the power fluid and the other pumping the well fluids. Another form of a hydraulic downhole pump that has become more popular is the jet pump, which converts the pressurized power fluid to a high-velocity jet that mixes directly with the well fluids. In the turbulent mixing, momentum and energy from the power fluid are added to the produced fluids.  The operating pressures in hydraulic pumping systems usually range from 2,000 to 4,000 psi. The most common pump used to generate this pressure on the surface is a multiplex positive displacement pump driven by an electric motor or multicylinder gas or diesel engine. Multistage centrifugal pumps and horizontal electrical submersible pumps (ESPs) have been used,  and some systems have been operated with the excess capacity in water-injection systems.  The hydraulic fluid usually comes from the well and can be either produced oil or water. A fluid reservoir at the surface provides surge capacity and is usually part of the cleaning system used to condition the well fluids for use as power fluid. Appropriate control valves and piping complete the system. A schematic of a typical hydraulic pumping system is shown in Fig. 14.1.
Types of Installations
The two basic types of installations are the "fixed"-pump and the "free"-pump design. In the fixed installation, the downhole pump is attached to the end of a tubing string and run into the well. Free-pump installations are designed to allow the downhole pump to be circulated into and out of the well inside the power-fluid string, or it can also be installed and retrieved by wireline operations.
Fixed-Pump InstallationsIn the fixed-insert (or tubing-conveyed) design, the pump typically lands on a seating-shoe in the larger tubing. Power fluid is normally directed down the inner tubing string, and the produced fluids and return power fluid flow to the surface inside the annulus between the two tubing strings, as shown in Part A of Fig. 14.2. These systems provide a passage for free gas in the annular space between the outer tubing string and the inside of the well casing, but to take full advantage of this gas-venting passage, the pump should be set below the perforations. The power-fluid string is usually ¾-in., 1-in., or 1¼-in. nominal tubing or 1-in., 1¼-in. or 1½-in. coiled tubing. The fixed-pump system is used mainly to fit a large-diameter downhole pump into restricted casing sizes and still retain the gas-vent feature. It also can be used to lift one or both zones of a dual well with parallel strings.
In the fixed-casing design, the tubing with the pump attached to its lower end is seated on a packer, as shown in Part B of Fig.14.2. With this configuration, the power fluid is directed down the tubing string, and the mixed power fluid and the produced well fluids return to the surface in the tubing/casing annulus. Because the well fluids enter the pump from below a packer, the pump must handle all the free gas. This type of installation is normally used with large-diameter high-capacity pumps in wells with little free gas, and if space permits, a gas-vent string can be run from below the packer to the surface. As with the fixed-insert design, this installation is no longer common, and both have been largely supplanted by the various free-pump installations. Note that in both of the fixed-type installations, when using a reciprocating piston pump, the power fluid mixes with the produced fluid after passing through the pump.
Free-Pump InstallationsThe free-pump feature is one of the most significant advantages of hydraulic pumping systems. Free-pump installations permit circulating the pump to the bottom, producing the well, and circulating the pump back to the surface for repair or size change. Fig. 14.3 shows pump in-and-out operations for a typical free-pump installation. They require that a bottomhole assembly (BHA) be run in on the tubing string. The BHA consists of a seating shoe and one or more sealbores above it and serves as a receptacle for the pump itself. BHAs are of robust construction and use corrosion-resistant sealing bores to ensure a long life in the downhole environmental conditions. The extensions needed on the BHA also can be adapted with different metallurgy to accommodate a changing environment. Once run in on the tubing string, the BHA normally remains in place for years, even though the downhole pump may be circulated in and out numerous times for repair or resizing. As shown in Fig. 14.4, a wireline-retrievable standing valve is landed in the seating shoe below the pump. The pump is run in the hole by placing it in the power-fluid string and pumping fluid down the tubing. When the pump reaches bottom, it enters the sealbores, begins stroking or jetting, and opens the standing valve. During normal pumping, this valve is held open by well fluid drawn into the pump suction. During pump-out, the normal flow of fluids is reversed at the surface with appropriate valving and pressure applied to the discharge flow path of the pump. This reversal of flow closes the standing valve and permits the pump to be circulated to the surface—a process that normally takes 30 minutes to 2 hours, depending on depth, tubing size, and the circulating flow rate.
The benefits of being able to circulate the downhole pump in and out of the well include reduced downtime and the ability to operate without a pulling unit for tubing, cable, or rod removal. Another significant advantage is that pressure and temperature recorders can be mounted on the pump to monitor downhole conditions with different pumping rates. At the conclusion of the test, circulating the pump to the surface also retrieves the recorder. Substituting a dummy pump for the normal production unit can be used to check for leakage of tubing pressure. Steaming, acidizing, or chemical treatment of the formation can be done if the pump is circulated out and the standing valve retrieved on wireline. A flow-through blanking tool may be run instead of the pump for such treatment if isolation of the power fluid and discharge flow paths is desired.
The casing-free installation, shown in Part C of Fig. 14.2, is attractive from an initial cost standpoint because it uses only one string of tubing. At first glance, it seems to be the same as the fixed-casing design, but the crucial difference is that instead of being attached to the end of the power-fluid string, the pump fits inside it to allow circulation into and out of the well. For a given diameter pump, this requires a larger-diameter string that reduces the annular flow path for the discharge fluids, but in most cases, a more than adequate flow area remains. Nominal tubing as small as 1½ in. can be run in systems with 2 7/8-in.-outside-diameter (OD) tubing used as casing, and coiled tubing as small as 1¼ in. can be run in systems with 2 3/8-in.-OD tubing used as casing. In the 1½-in. and 1¼-in. nominal-size tubing, only the jet pump can be used, while in 2 3/8-in.-OD or larger tubing, either jet or reciprocating pumps are suitable. Usually, 2 3/8-in.-OD power-fluid tubing is used in 4½-in.-OD or larger casing, 2 7/8-in.-OD tubing in 5½-in.-OD casing or larger, and 3½-in.-OD tubing in 6 7/8-in.-OD casing or larger. Because the BHA sits on a packer, the pump must handle all the gas from the well in addition to the liquids, even though a gas-vent string can be run if gas interference limits pump performance. In both the vented and unvented systems, the power fluid mixes with the produced fluids and returns to the surface. In wells where the produced fluid should be kept off the casing wall or where gas venting is desired, the parallel-free installation should be considered. This installation, which requires two parallel tubing strings, normally does not require a packer. As shown in Part D of Fig. 14.2, the BHA is suspended on the power-fluid string, and the return is either screwed into the BHA or is run separately with a landing spear that enters a bowl above the BHA. The tubing/casing annulus serves as a gas vent passage, and to take full advantage of this, the unit should be set below the perforations. In wells with corrosive gas and/or liquid, it may be undesirable to use the casing for return of gas or to have the liquid in the casing annulus. In such cases, a packer can be installed; however, the pump must handle all the gas and produced liquids.
Open and Closed Power-Fluid Systems
All installations discussed so far are open power-fluid (OPF) types, which means that all the power fluid and the produced fluid are mixed together after leaving the downhole pump and return to the surface together in a common flow passage. Jet pumps are inherently OPF pumps because the energy transfer depends on mixing the power fluid with the produced fluid. All reciprocating piston pumps (not jets) keep the power and produced fluids separate during the energy transfer process because there is a separate piston for each fluid. If the BHA has appropriate sealbores and passages to keep the two fluids separated, the power fluid can return to the surface in a separate tubing string, thus creating a closed power-fluid system.
Reverse-Flow SystemsConsiderations for a reverse-flow system for a jet-pump installation are the need to keep produced fluid off the casing, help minimize fluid friction losses, and aid in drillstem testing or unloading of wells. A reverse-flow installation is shown in Fig. 14.5. It uses the tubing/casing annulus for power fluid and the tubing string, which contains the pump, and is used for the combined power fluid and production. This protects the casing with inhibited power fluid and is most useful when severe corrosion is anticipated. In permanent installations, heavy wall casing should be a consideration to avoid casing burst conditions when power-fluid pressure is applied. In reverse-flow installations, the pump is run and retrieved on wireline in most cases but can be pumped in and out with a pusher-type locomotive.
Hydraulic pumps lend themselves to solution of the complex problem of the production of two separate zones or reservoirs in a single wellbore. To meet the artificial-lift requirements of the two distinct zones, two downhole pumps are usually required. It would be highly unusual if the same power-fluid pressure and rate were required for each zone; consequently, a separate power-fluid line for each pump is usually required. A number of completion configurations are possible, but small casing sizes and high gas/liquid ratios may severely hinder dual-well operation.
Principles of Operation
Reciprocating PumpsThe pump end of a hydraulic downhole pump is similar to a sucker-rod pump because it uses a rod-actuated plunger (also called the pump piston) and two or more check valves. The pump can be either single-acting or double-acting. A single-acting pump closely follows rod-pump design practices and is called single-acting because it displaces fluid on either the upstroke or downstroke (but not both). An example is shown schematically in Fig. 14.6. Fig. 14.7 shows a double-acting pump that has suction and discharge valves for both sides of the plunger, which enables it to displace fluids to the surface on both the upstroke and downstroke. With either system, motion of the plunger away from a suction valve lowers the pressure that holds the valve closed; it opens as the pressure drops, and well fluids are allowed to enter the barrel or cylinder. At the end of the stroke, the plunger reverses, forcing the suction valve to close and opening the discharge valving.
In a sucker-rod installation, the rod that actuates the pump plunger extends to the surface of the well and connects to the pumping unit; however, in hydraulic pumps, the rod is quite short and extends only to the engine pistons. The engine piston is constructed similarly to the pump plunger and is exposed to the power-fluid supply that is under the control of the engine valve. The engine valve reverses the flow of the power fluid on alternate half-strokes and causes the engine piston to reciprocate back and forth. Four-way engine valves are used with engines that switch from high-pressure to low-pressure power-fluid exhaust on both sides of the engine piston in an alternate manner. These engine (or reversing) valves are used with double-acting pump ends to give equal force on both upstroke and downstroke. Three-way engine valves are used with unequal-area engine pistons that always have high-pressure power fluid on one side and switch the power-fluid from high to low pressure on the other face of the piston. This type of reversing valve is commonly used with single-acting pumps that do not require a high force on the half-stroke because it is not displacing produced fluid to the surface. An example of this type of engine attached to a single-acting pump is illustrated in Fig. 14.8.
The engine or reversing valve can be activated by several methods. Commonly, ports on a rod direct pressure to control the engine valve at the extremes of the upstroke and downstroke, causing the valve to shift hydraulically. The shifting of the engine valve redirects the flow of power fluid to the engine piston and causes the reversal of the rod-and-plunger system. Alternatively, the engine can be mechanically "bumped" from one position to the other by the rod and plunger system as it nears the end of the upstroke and downstroke. Combinations of mechanical and hydraulic shifting are possible, and the engine valve may be located above the rod-and-plunger system, in the middle of the pump, or in the engine piston.
Note that the two designs illustrated and discussed do not exhaust the design possibilities offered by the various pump manufacturers; many combinations are possible, and tandem pumps or engine sections are sometimes advantageous. Examples of combinations of these design concepts can be seen in the cross-section schematics of the various pump types that accompany the pump specifications in Tables 14.1 through 14.3, which show the producing abilities and other factors that should be considered in designing reciprocating pumps. In the past, many tables were used in choosing the correct pump for the application, but today, the use of computers eliminates the errors inherent in reading charts and tables, making the process much simpler. Common to all designs, however, is the concept of a reversing valve that causes an engine piston (or pistons) to reciprocate back and forth, stroking the pump plunger (or plungers) that lifts fluid from the well.
Because the engine and pump are closely coupled into one unit, the stroke length can be controlled accurately; thus, the unswept area or clearance volume at each end of the stroke can be kept very small, leading to high compression ratios. This is very important when gas is present, and it generally prevents gas locking in hydraulic pumps. The engine valves and their switching mechanisms usually include controls to provide a smooth reversal and to limit the plunger speed under unloaded conditions. The unloaded plunger speed control is often called "governing" and minimizes fluid pound when the pump is not fully loaded with liquid. In this way, shock loads in the pump, as well as water hammer in the tubing strings, are softened, which reduces stress and increases life. (Caution: high pump speeds, at or above the rating, may significantly shorten piston pump run lives.)
Jet PumpsJet pumps are a type of downhole pump that can be used in hydraulic pumping systems instead of the reciprocating piston pumps previously discussed. They can be adapted to fit interchangeably into the BHAs designed for the stroking pumps. In addition, special BHAs have been designed for jet pumps to take advantage of their short length and their high-volume characteristics. Because of their unique characteristics under different pumping conditions, jet pumps should be considered as an alternative to the conventional stroking pumps.
Although technical references to jet pumps can be found as far back as 1852,  it was not until 1933 that a consistent mathematical representation was published that included suggestions for pumping oil wells.  Angier and Crocker applied for a patent on an oilwell jet pump in 1864 that looked very much like those currently marketed. Jacuzzi received a patent in 1930 for jet pumps that were subsequently used in shallow water wells successfully. McMahon also received the first of six patents on oilwell jet pumps in 1930. Apparently McMahon built and marketed pumps in California in the late 1930s, but they did not achieve widespread use. Hardware improvements and the advent of computer models for correct applications sizing in oil wells led to the successful marketing of jet pumps in 1970, and the use of jet pumps has grown steadily since then. More recent publications on hydraulic pumping that describe the use of jet pumps in oil wells include those by Wilson, Bell and Spisak, Christ and Zublin, Nelson,  Brown,  Clark,  Bleakley,  and Petrie et al.  Much of the following discussion derives from Refs. 15, 18, and 19 .
An example of the simplest downhole jet free-pump completion, the single-seal style, is shown in Fig. 14.9. The most significant feature of this device is that it has no moving parts; the pumping action is achieved through energy transfer between two moving streams of fluid. The high-pressure power fluid, supplied from the surface, passes through the nozzle, where its potential energy (pressure) is converted to kinetic energy in the form of a very-high-velocity jet of fluid. Well fluids surround the power-fluid jet at the tip of the nozzle, which is spaced back from the entrance of the mixing tube. The mixing tube, usually called the throat, is a straight, cylindrical bore about seven diameters long with a smoothed radius at the entrance. The diameter of the throat is always larger than the diameter of the nozzle exit, allowing the well fluids to flow around the power-fluid jet and be entrained by it into the throat. In the throat, the power fluid and produced fluid mix, and momentum is transferred from the power fluid to the produced fluid, causing its energy to rise. By the end of the throat, the two fluids are intimately mixed, but they are still at a high velocity, and the mixture contains significant kinetic energy. The mixed fluid enters an expanding area diffuser that converts the remaining kinetic energy to static pressure by slowing down the fluid velocity. The pressure in the fluid is now sufficiently high to flow it to the surface from the downhole pump.
With no moving parts, jet pumps are rugged and tolerant of corrosive and abrasive well fluids. The nozzle and throat are usually constructed of tungsten carbide or ceramic materials for long life. Successful jet-pump adaptations have also been made for sliding side doors in both the normal and reverse-flow configurations. These are normally run in on wireline or as a fixed or conventional installation on continuous coiled tubing and have been successful in offshore drillstem testing (DST) of heavy-crude reservoirs. Other applications include the dewatering of gas wells. 
With different sizes of nozzles and throats, jet pumps can produce wells at less than 50 B/D or in excess of 15,000 B/D. To achieve high rates, a special BHA is required as the BHA itself is used as a crossover for the production, allowing for larger passages for the produced fluid to travel to the jet nozzle as shown in Fig. 14.10. As with all hydraulic pumping systems, a considerable range of production is possible from a particular downhole pump by controlling the power-fluid supply at the surface, but for any given size of tubing, the maximum achievable rates are usually much higher than those possible with stroking pumps. Significant free-gas volumes can be handled without the problems of pounding or excessive wear associated with positive-displacement pumps, or the inlet choking encountered in centrifugal pumps. The lack of vibration and the free-pump feature make them ideal for use with pumpdown pressure recorders to monitor BHPs at different flow rates.
Because they are high-velocity mixing devices, there is significant turbulence and friction within the pump, leading to lower horsepower efficiencies than achieved with positive-displacement pumps. This often leads to higher surface horsepower requirements, although some gassy wells may actually require less pressure. Jet pumps are prone to cavitation at the entrance of the throat at low pump intake pressures, and this must be considered in design calculations. Also, because of the nature of their performance curves, the calculations used for installation design are complex and iterative in nature and are best handled by computers. Their overall energy efficiencies are low, which may lead to high energy costs; despite these limitations, their reliability, low maintenance costs, and volume capability make them attractive in many wells, and their use has increased since commercial introduction in the early 1970s.
Performance CharacteristicsIntuitively, larger-diameter nozzles and throats would seem to have higher flow capacities, and this is normally the case. The ratio of the nozzle area to the throat area is an important variable because it determines the trade-off between produced head and flow rate. Fig. 14.11 shows a schematic of the working section of a jet pump. If, for a given nozzle, a throat is selected such that the area of the nozzle, An, is 60% of the area of the throat, At, a relatively high-head, low-flow pump will result. There is a comparatively small area, As, around the jet for well fluids to enter. This leads to low production rates compared to the power-fluid rate, and because the energy of the nozzle is transferred to a small amount of production, high heads develop. Such a pump is suited for deep wells with high lifts, and substantial production rates can be achieved if the pump is physically large, but the production rate will always be less than the power-fluid rate.
If a throat is selected such that the area of the nozzle is only 20% of the area of the throat, much more flow area around the jet is available for the production. However, because the nozzle energy is transferred to a large amount of production compared to the power-fluid rate, lower heads will be developed. Shallow wells with low lifts are candidates for such a pump.
Any number of such area combinations is possible to match different flow and lift requirements. Attempting to produce small amounts of production compared to the power-fluid rate with nozzle/throat-area ratio of 20% will be inefficient because of high-turbulent mixing losses between the high-velocity jet and the slow-moving production. Conversely, attempting to produce high production rates compared to the power-fluid rate with a nozzle/throat-area ratio of 60% will be inefficient because of high friction losses as the produced fluid moves rapidly through the relatively small throat. Optimal ratio selection involves a trade-off between these mixing and friction losses.
As a type of dynamic pump, the jet pump has characteristic performance curves similar to those of an ESP. A family of performance curves is possible, depending on the nozzle pressure supplied to the pump from the surface. Different sizes of throats used in conjunction with a given nozzle size give different performance curves. The curves are generally fairly flat, especially with the larger throats, which makes the jet pump sensitive to changes in intake or discharge pressure. Because variable fluid mixture densities, gas/liquid ratios, and viscosity affect the pressures encountered by the pump, the calculations to simulate performance are complex and iterative in nature and lend themselves to a computer solution.
Cavitation in Jet Pumps
Because the production must accelerate to a fairly high velocity (200 to 300 ft/sec) to enter the throat, cavitation is a potential problem. The throat and nozzle flow areas define an annular flow passage at the entrance of the throat. The smaller this area is, the higher the velocity is of a given amount of produced fluid passing through it. The static pressure of the fluid drops as the square of the velocity increases, declining to the vapor pressure of the fluid at high velocities. This low pressure causes vapor cavities to form, a process called cavitation. This results in choked flow into the throat, and production increases are not possible at that pump-intake pressure, even if the power-fluid rate and pressure are increased. Subsequent collapse of the vapor cavities, as pressure is built up in the pump, may cause erosion known as cavitation damage. Thus, for a given production flow rate and pump intake pressure, there is a minimum annular flow area required to keep the velocity low enough to avoid cavitation. This phenomenon has been the subject of numerous investigations—the most notable being that of Cunningham and Brown,  who used actual oilwell pump designs at the high pressures used in deep wells.
The description of the cavitation phenomenon suggests that if the production flow rate approaches zero, the potential for cavitation will disappear because the fluid velocities are very low. Under these conditions, however, the velocity difference between the power-fluid jet and the slow-moving production is at a maximum, which creates an intense shear zone on the boundary between them, generating vortices, the cores of which are at a reduced pressure. Vapor cavities may form in the vortex cores, leading to erosion of the throat walls as the bubbles collapse because of vortex decay and pressure rises in the pump. Although no theoretical treatments of this phenomenon have been published, it has been the subject of experimental work, which has led to the inclusion, by suppliers, of potential damage zones on their published performance prediction plots. This experimental correlation predicts cavitation damage at low flow rates and low pump-intake pressures before the choked flow condition occurs. Field experience has shown, however, that in most real oil wells, the erosion rate in this operating region is very low, probably because of produced gas cushioning the system by reducing the propagation velocity of the bubble-collapse shock waves. It is generally agreed that this phenomenon is of concern only in very-high-water-cut wells with virtually no gas present. Under these conditions, cavitation erosion has been observed even at very low production rates; however, if a jet pump is operated near its best efficiency point, the shear vortices are a distinctly second-order effect in the cavitation process.
Nozzle and Throat SizesEach manufacturer has different sizes and combinations of nozzles and throats. Manufacturers A and B increase the areas of nozzles and throats in a geometric progression (i.e., the flow area of any nozzle or throat is a constant multiple of the area of the next smaller size). Manufacturer B’s factor is 1.29155, and Manufacturer A’s factor is 4/π = 1.27324. The system of sizes offered by Manufacturer C uses a similar geometric progression concept but does not use the same factor over the total range. In the smaller sizes, where the change in horsepower per size is small, the rate of increase in area is more rapid than in the systems of Manufacturers A and B. In the larger, higher-horsepower sizes, the percent increase in size is less rapid than in the other systems to limit the incremental increase in horsepower. The sizes offered by Manufacturer C cover a slightly larger range than those of Manufacturers A and B. The sizes from these manufacturers are listed in Table 14.4. The maximum sizes of nozzles and throats that are practical in pumps for a given tubing size depend on the fluid passages of the particular pump, BHA, swab nose, and standing valve. Single-seal pumps cannot use nozzles as large as those practical in higher-flow, multiple-seal pumps. In general, nozzles larger than 0.035 in.2 in flow area are used only in pumps for 2½- and 3½-in. tubing.
The strict progression used by Manufacturers A and B establishes fixed area ratios between the nozzles and different throats. A given nozzle matched with the same number throat always gives the same area ratio: 0.380 in Manufactures A’s system and 0.400 in Manufacturer B’s system (Table 14.4). This is called the A ratio. Successively larger throats matched with a given nozzle give the B, C, D, and E ratios. In the systems of Manufacturers A and B, the nozzle size and ratio designate the size of a pump. Examples are 11-B, which is a No. 11 nozzle and a No. 12 throat, and 6-A, which is a No. 6 nozzle and a No. 6 throat.
Because the size progression for the nozzles and throats in Manufacturer C’s system is not constant over the whole range, the nozzle/throat combinations do not yield fixed ratios. However, the ratios that result cover the same basic range as the other two systems. The actual ratios are listed in Table 14.5. In Manufacturer C’s system, the nozzle and mixing tube (throat) sizes designate the size of a pump. An example is C-5, which are the size C nozzle and the No. 5 throat. This combination has an area ratio of 0.32. The annular flow areas of Manufacturer C’s jet pumps used in cavitation calculations are also included in Table 14.6. The annular areas for Manufacturers A and B’s jet pumps are listed in Tables 14.6 and 14.7.
The most commonly used area ratios fall between 0.235 and 0.400. Area ratios greater than 0.400 are sometimes used in very deep wells with high lifts or when only very low surface operating pressures are available and a high head regain is necessary. Area ratios less than 0.235 are used in shallow wells or when very low BHPs require a large annular flow passage to avoid cavitation. The smaller area ratios develop less head but may produce more fluid than is used for power fluid (FmfD > 1.0). Where the curves for different area ratios cross, the ratios have equal production and efficiency; however, different annular flow areas (As) may give them different cavitation characteristics.
Jet-Pump Application SizingThe widespread current use of jet pumps can be credited to the advent of computer programs capable of making the iterative calculations necessary for application design. Jet-pump performance depends largely on the pump discharge pressure, which in turn is strongly influenced by the gas/liquid ratio, FgL; in the return column to the surface, higher values of FgL lead to reduced pump discharge pressure. Because the jet pump is inherently an OPF device, FgL depends on the formation gas/oil ratio (GOR) and on the amount of power-fluid mixed with the production, which in turn depends on the size of the nozzle and the operating pressure. As the power-fluid pressure is increased, the lift capability of the pump increases, but the additional power-fluid rate decreases FgL, thereby increasing the effective lift. Finding a match between the power-fluid rate, the pump performance curve and the pump discharge pressure, p, is an iterative procedure involving successive refined guesses.
The various suppliers of jet pumps also have developed in-house computer programs for application design that are faster than the past calculator routines and incorporate more correlation for fluid properties and the pump discharge pressure. The object of the calculation sequence is to superimpose a jet-pump performance curve on the inflow performance relationship (IPR) curve of the well and to note the intersections that represent the pump performance in that particular well. Therefore, a plot of the best estimate of the IPR or productivity index (PI) curve of the well is the starting point. An example of a completed performance plot in this format is shown in Fig. 14.12.
Calculation Sequence and Supplemental EquationsFig. 14.13 shows a typical jet-pump installation with the appropriate pressures that determine pump operation. Although a parallel installation is shown for clarity of nomenclature, the same relationships hold for the casing-type installation.
Downhole Pump Accessories
Swab CupsA number of accessories are available for downhole pumping systems. Free-pump systems require swab cups and a standing valve to accomplish the pump-in and pump-out operations. The swab cups are carried on a mandrel, extending above the pump, which may contain a check valve to limit the amount of fluid by passing the pump as it is circulated to the surface. If the pump does not enter a lubricator on the wellhead, the check valve may include a valve bypass that is actuated when the pump enters the wellhead catcher to prevent excessive pressure buildup. Two examples of swab cup assemblies are shown in Fig. 14.14. Jet pumps usually use the simpler system.
Standing ValvesStanding valves are necessary in free-pump systems to create a "U" tube and prevent the circulating fluid from flowing back into the reservoir. During pumping operations, the standing valve is opened by flow from the formation to the pump suction; whenever the pump is shut down, the standing valve closes. In some cases, the standing-valve ball is held open by a small magnet to prevent it from cycling during reciprocating pump-stroking reversals. When the downhole pump is unseated, fluids attempting to flow back into the formation wash the ball off the magnet and onto the seat. The standing valve is wireline-retrievable and includes a provision for draining the tubing before attempting to pull it. In most cases, the standing valve forms the no-go and bottom seal for the pump. Some jet-pump installations, however, use high-flow designs that do not serve as a pump seat. An example of each type is shown in Fig. 14.15.
To obtain producing BHPs at several different withdrawal rates, downhole pressure recorders are often run in conjunction with hydraulic pumps, hung below the standing valve. While this arrangement provides not only pressure drawdown but also pressure-buildup data, it has the disadvantage of requiring wireline operations to run and retrieve the recorder. Some reciprocating pumps can be run with a pressure recorder attached, which eliminates the wireline operations but does not permit observation of pressure buildup because the recorder is above the standing valve. Virtually all jet pumps can be run with recorders attached, and very smooth recordings are obtained.
Dummy pumps are sometimes run to blank off one or more tubing strings so that they may be checked for leaks. If the dummy pump has a fluid passage in it, the terms "flow-through dummy" or "blanking tool" are often used. These tools are useful for acidizing or steaming.
Screens and Filters
To protect the downhole pump from trash in the well, various types of screens and filters are sometimes run. Because circulating pumps in and out of a well may dislodge scale and corrosion products in the tubing, a starting filter can be attached to the swab-cup assembly to filter the power fluid. Because this must be a relatively small filter, it will eventually plug up, and an automatic bypass arrangement is provided. This system collects foreign material during the crucial startup phase with a newly installed pump. For long-term operation, power-fluid and pump intake screens or strainers are used, which exclude large-diameter objects that could damage or plug the pump.
Safety ValvesIn some areas, subsurface safety valves are required. When a packer is set and the BHA is above it, a wireline-retrievable safety valve can be installed between the standing valve and the packer to isolate the formation. The safety valve is normally closed unless the pump supplies high-pressure fluid to it by way of a control line run from the main power-fluid tubing just above. The pump discharge pressure provides the reference pressure to the safety valve. When the pump is on bottom and power-fluid pressure is applied to it, the safety valve opens to allow well fluid to enter the pump. Most safety valves will not hold pressure from above, so the standing valve is still necessary for circulating the pump in and out of the well. Fig. 14.16 illustrates this type of installation.
Surface PumpsHydraulic pumping systems have evolved toward the use of relatively high pressures and low flow rates to reduce friction losses and to increase the lift capability and efficiency of the system. Surface operating pressures are generally between 2,000 and 4,000 psi, with the higher pressures used in deeper wells, and power-fluid rates may range from a few hundred to more than 3,000 B/D. While some surface multistage centrifugal pumps are rated to this pressure range, they are generally quite inefficient at the modest flow rates associated with single-well applications. Multistage centrifugals can be used effectively when multiple wells are pumped from a central location. The surface pump for a single well or for just a few wells must be a high-head and low-specific-speed pump. Wide experience in the overall pumping industry has led to the use of positive-displacement pumps for this type of application, and triplex or quintuplex pumps, driven by gas engines or electric motors, power the vast majority of hydraulic pump installations. See Fig. 14.17.
Multiplex pumps consist of a power end and a fluid end. The power end houses a crankshaft in a crankcase. The connecting rods are similar to those in internal combustion engines, but connect to crossheads instead of pistons. The fluid end houses individual plungers, each with intake and discharge check valves usually spring loaded, and is attached to the power end by the spacer block, which houses the intermediate rods and provides a working space for access to the plunger system. Most units being installed in the oil field are of the horizontal configuration, which minimizes contamination of the crankcase oil with leakage from the fluid end. Vertical installations are still found, however, particularly with oil as the pumped fluid or when space is at a premium, as in townsite leases.
Multiplex pumps applied to hydraulic pumping usually have stroke lengths from 2 to 7 in. and plunger diameters between 1 and 2½ in. The larger plungers provide higher flow rates but are generally rated at lower maximum pressure because of crankshaft loading limitations. The larger plungers provide higher flow rates, but are generally rated at lower maximum pressure because of crankshaft loading limitations. The normal maximum rating of multiplexes for continuous duty in hydraulic pumping applications is 5,000 psi, with lower ratings for the larger plungers, but applications above 4,000 psi are uncommon. Multiplex pumps are run at low speed to minimize vibration and wear and to avoid dynamic problems with the spring-loaded intake and discharge valves. Most applications fall between 200 and 450 rev/min, and because this is below the speeds of gas engines or electric motors, some form of speed reduction is usually required. Belt drives are found on some units, although gear reduction is more common while gear-reduction units are integral to some multiplexes and separate on others. A variety of reduction ratios are offered for each series of pumps. Because a positive-displacement pump has an essentially constant discharge flow rate for a given prime-mover speed, bypass of excess fluid normally is used to match a particular pressure and flow demand. Another option that has been used successfully is to drive the multiplex pump through a four-speed transmission, which greatly enhances the flexibility of the system. This allows much closer tailoring of the triplex output to the demand, thereby pumping at reduced speed when needed, which also tends to increase the life of such components as the packing and valving.
Each plunger pumps individually from a common intake manifold into a common discharge, and because discharge occurs only on the upstroke, there is some pulsation, for which pulsation dampers are commonly used.
Two types of plunger systems are in common use. For oil service, a simple and effective plunger-and-liner system is used that consists of a closely fitted metallic plunger inside a metallic liner. Sprayed metal coatings or other hard-facing means are often used to extend the life of the plunger and liner. When pumping water, the metal-to-metal system is not practical because the fit would have to be extremely close to keep leakage to an acceptable level. Galling and scoring are problems with close fits and the low lubricity of water, and to solve this problem, spring-loaded packing systems are used that do not require adjusting. The advent of high-strength aramid fibers for packing, in conjunction with other compounds to improve the friction characteristics, has resulted in a pronounced improvement in the ability of the pump to handle high-pressure water for extended periods of time. Water still presents a more severe challenge than oil, however, and water systems show much better life if operated at or below 3,500 psi.
Suction conditions are important to multiplex operation. Friction losses in piping, fluid end porting, and across the suction valving reduce the pressure available to fill the pumping chamber on the plunger downstroke, and if these losses are sufficiently great, cavitation may result. When pumping oil with dissolved gas, the reduction in pressure liberates free gas and causes knocking, so it is necessary to have a positive head on the suction side to overcome the friction losses. In addition, another phenomenon known as "acceleration head" must be considered. The flow in the suction piping must accelerate and decelerate a number of times for each crankshaft revolution. For the fluid (which has inertia) to follow the acceleration, energy must be supplied, which is then returned to the fluid on deceleration. The energy supplied during acceleration comes from a reduction in the pressure in the fluid, and if this drops too low, cavitation or gas liberation will result. The minimum suction head for the multiplex pump is then the sum of the friction losses and the acceleration head. Although the pump can draw a vacuum, this will flash gas and may tend to suck air across the valve or plunger packing. Manufacturers of multiplex pumps recommend appropriate suction charging pressures for their products, but it is worth noting that long, small-diameter suction lines increase the acceleration head loss and friction loss. It is therefore recommended for suction lines to be short and of large diameter, with no high spots to trap air or gas. Suction stabilizers or pulsation dampeners that tend to absorb the pulsations from the pump also reduce acceleration head, and users are encouraged to follow good piping practices in the installation of surface pumps.
In many cases, sufficient hydrostatic head is not available to provide the necessary suction pressure, and charge pumps are used to overcome this problem. Positive displacement pumps of the vane or crescent-gear type driven from the triplex have been used extensively, but they require a pressure-control valve to bypass excess fluid and match the multiplex displacement. Where electric power is available, centrifugal charge pumps have given excellent service. Centrifugal pumps generally need to run at speeds considerably above the multiplex speed, and so driving them from the multiplex presents problems, particularly with a gas engine drive where prime-mover speed variations cause significant variations in the charge-pump output pressure.
While good charging pressures are necessary to ensure proper loading and smooth operation, there are problems associated with very high charge pressures. These add to the crankshaft loading, and for charge pressures above about 250 psi, it is advisable to derate the maximum discharge pressure by one third of the charge pressure. High charge pressures also can adversely affect the lubrication of bearings, particularly in the crosshead wristpin. In addition, the mechanical efficiency of multiplex pumps is some 3 to 5% lower on the suction side compared to the discharge side.  Consequently, the combination of a charge pump and multiplex pump is most efficient with low charging pressures and a high boost by the multiplex pump. The charging pressure should therefore be limited to that necessary to give complete filling of the multiplex pump with a moderate safety allowance for variations in the system parameters.
In some cases, it is desirable to inject corrosion inhibitors or lubricants into the multiplex suction, and fresh water is sometimes injected to dissolve high salt concentrations. In severe pumping applications with low-lubricity fluids, lubricating oil is sometimes injected or dripped onto the plungers in the spacer block area to improve plunger life. Injection pumps are often driven from the multiplex drive for these applications. A troubleshooting guide for multiplex pumps is given in Table 14.8.
Fluid ControlsVarious types of valves are used to regulate and to distribute the power-fluid supply to one or more wellheads. Common to all free-pump systems is a four-way valve or wellhead control valve, which is mounted at the wellhead, as shown in Fig. 14.18. Its function is to provide for different modes of operation by shifting it to different positions. To circulate the pump into the hole, as shown in Fig. 14.15, power fluid is directed down the main tubing string. The power fluid begins to operate the pump once it is on bottom and seated on the standing valve. In the pump-out mode, power fluid is directed down the return tubing or casing annulus to unseat the pump and to circulate it to the surface. When the pump is on the surface, putting the valve in the bypass and bleed position permits the well to be bled down and the pump to be removed and replaced.
Most systems include a constant-pressure controller, as shown in Fig. 14.19, which maintains a discharge-pressure load on the multiplex pump by continuously bypassing the excess discharge fluid. It generally operates on the principle of an adjustable spring force on a piston-and-valve assembly that is pressure compensated. If the pressure rises on the high-pressure side, which is being controlled because of changing system loads, the pressure forces on the various areas within the valve will cause the valve to open and to bypass more fluid, restoring the high-pressure side to the preset condition. Jet pumps frequently are operated with a constant-pressure valve as the only surface control valve. The constant-pressure controller can be used to regulate the pressure on a manifold assembly serving multiple wells.
Reciprocating downhole pumps are usually regulated with a constant-flow control valve. The downhole unit can be maintained at a constant stroking rate if a constant volume of power fluid is supplied to it, and the constant-flow control valve is designed to provide a preset flow rate even if the downhole operating pressure fluctuated because of changing well conditions. Because this valve does not bypass fluid, it must be used with a constant-pressure controller on the higher-pressure or inlet side.
Where a number of wells are to be pumped from a central battery, a control manifold is used to direct the flows to and from the individual wells. Control manifolds are designed to be built up in modular fashion to match the number of wells being pumped and are generally rated for 5,000 psi working pressure. A constant pressure control valve regulates the pressure on the common power fluid side of the manifold. This pressure is generally a few hundred pounds per square inch greater than the highest pressure demanded by any well to allow proper operation of the individual well-control valves. Individual constant-flow control valves regulate the amount of power fluid going to each well. The use of a constant pressure valve allows excess fluid to bypass at the highest pressure. Meter loops or individual meters for each station can be integrated into the manifold.
Some wells flow or "kick back" when the operator is attempting to remove or insert a pump into the wellhead. Also, the presence of water may make it inadvisable to open up the entire tubing string for pump insertion and removal. The use of a lubricator allows the master valve below the wellhead to be closed, and the entire lubricator with the pump in it to be removed from the wellhead. The lubricator is essentially an extended piece of the tubing with a sideline to allow fluid flow when the pump is circulating in or out of the hole.
The function of the surface treating systems is to provide a constant supply of suitable power fluid to be used to operate the subsurface production units. The successful and economical operation of any hydraulic pumping system is to a large extent dependent on the effectiveness of the treating system in supplying high-quality power fluid. The presence of gas, solids, or abrasive materials in the power fluid adversely affects the operation and wears both the surface and downhole units. Therefore, the primary objective in treating crude oil or water for use as power fluid is to make it as free of gas and solids as possible. In addition, chemical treatment of the power fluid may be beneficial to the life of the downhole unit. In tests, it has been found that for best operation of the unit, a maximum total solids of 20 ppm, maximum salt content of 12 lbm/1,000 bbl oil, and a maximum particle size of 15 μm should be maintained. (These norms were established using oil in 30 to 40°API gravity range). It has been observed, however, that acceptable performance has been achieved in many cases where these values were exceeded, especially with the use of jet pumps and larger nozzles and throats. When using piston hydraulic pumps in heavy crude, these limitations have been exceeded and satisfactory results achieved, probably because the resulting wear does not increase leakage to the same degree. The periodic analysis of power fluid indicates steps to be taken for improved operation. For example, if the power-fluid analysis shows that iron sulfide or sulfate compounds make up the bulk of the solids, then a corrosion or scale problem exists that would require the use of chemical inhibitors to correct the problem. Water is the primary power fluid being used for jet pumping on offshore platforms and in applications where the majority of produced fluid being made is water. Water requires that a lubricant be added for use with reciprocating pumps. Other considerations in the choice of water or oil as a power fluid include:
- Maintenance on surface pumps is usually less with the use of oil. The lower bulk modulus of oil also contributes to reduced pressure pulsations and vibrations that can affect all the surface equipment.
- Well testing for oil production is simpler with water as the power fluid because all the oil coming back is produced oil. With oil power fluid, the power rate must be closely metered and subtracted from the total oil returning to surface. This can be a source of considerable error in high-water-cut wells where the power oil rate is large compared to the net production.
- In high-friction systems, as sometime occurs with jet pumps in restricted tubulars, the lower viscosity of water can increase efficiency. With no moving parts, the jet pump is not adversely affected by the poor lubrication properties of water.
- In deep casing-type installations, particularly with a jet pump, water when used as the power fluid can load up in the casing annulus return, negating any beneficial gas lifting effects for the produced gas.
It has been found that, in most cases, an upward velocity of 1ft/hr is low enough to provide sufficient gravity separation of entrained particles to clean power fluid to requirements, provided that there is no free gas in the fluids or large thermal effects.
Open Power-Fluid SystemA typical power-oil treating system that has proved adequate for most OPF systems, when stock-tank-quality oil is supplied, is shown in Fig. 14.20. This system has the general characteristic that all return fluids from the well, both production and power fluid, must pass through the surface treating facility. The power-oil settling tank in this system is usually a 24-ft-high, three-ring, bolted steel tank. A tank of this height generally provides adequate head for gravity flow of oil from the tank to the multiplex pump suction. If more than one multiplex pump is required for the system, individual power-oil tanks can be set up for each pump, or a single large tank can be used, whichever is more economical and best meets the operating requirements. If a single large tank supplies the suction for several pumps, individual suction lines are preferable.
The gas boot is essentially a part of the power-oil tank; its purpose is to provide final gas/oil separation so that the oil will be stable at near-atmospheric pressure. If the gas is not sufficiently separated from the oil, entrained free gas can enter the power-oil tank and destroy the settling process by causing the fluid in the tank to roll. The following piping specifications for the gas boot are necessary to ensure undisturbed settling:
- The gas-boot inlet height should be 4 ft above the top of the power-oil tank to allow the incoming fluid to fall, so that the agitation will encourage gas/oil separation.
- The top section of the gas boot should be at least 3 ft in diameter and 8 ft higher than the top of the power-oil tank. These two factors will provide a reservoir that should absorb the volume of the surges.
- The gas line out of the top of the boot should be tied into the power-oil tank and stock-tank vent line with a riser on the top of the power-oil tank. In the event the gas boot does become overloaded and kicks fluid over through the gas line, this arrangement will prevent the raw or unsettled fluid from being dumped in the top of the power-oil tank, where it may contaminate the oil drawn off to the multiplex. A minimum diameter of 3 in. is recommended for the gas line.
- The line connecting the gas boot to the power-oil tank should be at least 4 in. in diameter. This is necessary to minimize restrictions to low during surge loading of the boot.
Oil entering a large tank (at the bottom and rising to be drawn off the top) tends to channel from the tank inlet to the outlet; thus, an inlet spreader is used. The purpose of the spreader is to reduce the velocity of the incoming fluid by distributing the incoming volume over a large area, thus allowing the fluid to rise upward at a more uniform rate. The recommended spreader consists of a round, flat plate with a diameter approximately half that of the tank with a 4-in. skirt that has 60° triangular, saw-tooth slots cut in it. The slots provide automatic opening adjustment for varying amounts of flow. It is essential that they be cut to uniform depth to obtain an even distribution of flow. This type of spreader must be installed with the tops of all the slots in a level plane to prevent fluid from "bumping out" under a high side, and it should be mounted about 2 ft above the bottom rim of the tank.
The location of the stock-tank take-off and level control is important because it establishes the effective settling interval of the power-oil tank and controls the fluid level. All fluid coming from the spreader rises to the stock take-off level, where stock-tank oil is drawn off. Fluid rising above this level is only that amount required to replace the fluid withdrawn by the multiplex pump, and it is in this region that the power-oil settling process takes place. The light solids settled out are carried with the production through the stock-tank take-off, and the heavier particles settle to the bottom, where they must be removed periodically. The location of the stock take-off point should be within 6 ft of the spreader. The height to which the stock oil must rise in the piping, to overflow into the stock tank, determines the fluid level in the power-oil tank. The diameter of the piping used should be sufficient to provide negligible resistance to the volume of flow required (4-in. minimum diameter recommended). The extension at the top of the level control is connected to the gas line to provide a vent that keeps oil in the power-oil tank from being siphoned down to the level of the top of the stock tank.
The power-oil outlet should be located on the opposite side of the power-oil tank from the stock take-off outlet to balance the flow distribution within the tank. Because the fluid level in the tank is maintained approximately 18 in. from the top of the tank, the location of the upper outlet, for use in starting up or filling tubing strings, depends on estimated emergency requirements and the capacity per foot of the tank. A distance of 7 ft from the top of the tank is usually sufficient. This lower outlet line contains a shutoff valve that is kept closed during normal operations so that the full settling interval is used.
Closed Power-Fluid Systems
In the closed power-fluid system, the power fluid returns to the surface in a separate conduit and need not go through the surface production treating facilities. The consequent reduction in surface treating facilities can tend to offset the additional downhole cost of the system. Virtually all closed power-fluid systems are in California because of the large number of townsite leases and offshore platforms, and water is usually the power fluid. Gravity settling separation in the power-fluid tank ensures that the power fluid remains clean despite the addition of solids from power-fluid makeup, corrosion products, and contamination during pump-in and pump-out operations. The power-fluid makeup is required to replace the small amount of fluid lost through fits and seals in the downhole pump and wellhead control valve. A certain amount of power fluid is also lost during circulating operations as well.
Single-Well SystemsThe central battery systems previously discussed have been used successfully for years and provide a number of benefits. The use of lease fluid treating facilities as part of the of the hydraulic system ensures good, low-pressure separation of the gas, oil, water, and solids phases present in any system. Good triplex charging of clean, gas-free oil and consistently clean power fluid supplied to the downhole pump are desirable features of this system. The lease treating facilities, however, must have sufficient capacity to process both the well production and the return power fluid. When the wells are closely spaced, the clustering of power generation, fluid treating, and control functions in one location (but sufficiently spread out) is very efficient and allows good use of the installed horsepower. Because the system is not limited by production variations on any one well, an adequate supply of the desired power fluid is ensured by the size of the system. A further benefit associated with the use of the lease separation facilities is the option of a closed power-fluid system. When well spacing is large, however, long, high-pressure power-fluid lines must be run. Also, individual well testing is complicated by the need to meter the power-fluid rate for each well, which can introduce measurement errors. As a final consideration, only a few wells in a field may be best suited to artificial lift by hydraulic pumping, so the installation of a central system is difficult to justify.
To address the limitations of the central battery system, single-well systems have been designed,  many of the requirements of which are the same as for a central battery. The oil, water, gas, and solids phases must be separated to provide a consistent source of power fluid to run the system. A choice of water or oil power fluid should be possible, and the fluid used as power fluid must be sufficiently clean to ensure reliable operation and that it is gas-free at the multiplex suction to prevent cavitation and partial fluid end loading. An adequate reservoir of fluid must be present to allow continuous operation and the various circulating functions associated with the free-pump procedures. Finally, a means of disposing of and measuring the well production to the lease treating and storage facilities must be provided.
To achieve these objectives, several of the manufacturers of hydraulic pumping units offer packaged single-well systems that include all the control, metering, and pumping equipment necessary. All components are skid mounted on one or two skids to facilitate installation at the well and to make the systems easily portable if the unit is to be moved to a different well. Usually, the only plumbing required at the wellsite is the power fluid and return-line hookup at the wellhead and the connection of the vessel outlet to the flowline.
An example of a typical single-well power unit is illustrated in Fig. 14.1. All units of this type share certain design concepts, with small variations depending on the manufacturer. Either one or two pressure vessels are located at the wellsite. The size of the main reservoir vessel depends on the nature of the well and the tubular completion. The reservoir vessel should ensure that, if the wellhead partially empties the return conduit to the flowline, adequate capacity remains to operate the downhole unit until production returns re-enter the vessel. Even if the well does not head, extra capacity is needed. When the unit is shut down for maintenance or pump changes, that portion of the return conduit occupied by gas needs to be filled from the vessel to unseat the pump and to circulate it to the surface. The vessel sizes normally used range from 42 × 120 in. to 60 × 240 in. In some wells, even the largest vessel may not be able to compensate fully for heading, in which case it is common to use backpressure to stabilize the heading. The vessels themselves are normally in the 175- to 240-psi working pressure range, with higher ratings available for special applications. Coal-tar-epoxy internal coatings are common, with special coatings for extreme cases.
The return power fluid and production for the well enter the vessel system where basic separation of water, oil, and gas phases takes place. Free gas at the vessel pressure is discharged to the flowline with a vent system that ensures a gas cap in the vessel at all times, while the oil and water separate in the vessel, and the desired fluid is withdrawn for use as power fluid. The power fluid passes through one or more cyclone desanders to remove solids before entering the multiplex suction, where it is pressurized for reinjection down the power-fluid tubing. Any excess multiplex output that is bypassed for downhole pump control is returned to the vessel. The underflow from the bottom of the cyclone desanders contains a high-solids concentration and is discharged either into the flowline or back into the vessel system. Once the system is stabilized on the selected power fluid, the well production of oil, water, and gas is discharged into the flowline from the vessel, which is maintained at a pressure above the flowline. Because the flowline is carrying only what the well makes, additional treating and separating facilities are not needed, as they are in the central battery system that handles mixed well production and power fluid. This feature also facilitates individual well testing.
Overall, simple gravity dump piping, which consists of a riser on the outside of the vessel, controls the fluid level in the vessel system. To prevent siphoning of the vessel, the gas-vent line is tied in the top of the riser as a siphon breaker. The choice of oil or water power fluid is made by selection of the appropriate take-off points on the vessel so that the production goes to the flowline and the power fluid goes to the multiplex pump. If the multiplex suction is low in the vessel and the flowline is high in the vessel, water will tend to accumulate in the vessel and will be the power fluid. If the multiplex suction is high in the vessel and the flowline is low, oil will tend to accumulate in the vessel and will be the power fluid. Opening and closing appropriate valves sets the system up for the chosen power fluid. The multiplex suction outlets are positioned with respect to the overall fluid level in the vessel to avoid drawing power fluid from the emulsion layer between the oil and water because this layer generally contains a significantly higher concentration of solids and is not easily cleaned in the cyclones.
The power-fluid cleaning is accomplished with cyclone desanders that require a pressure differential across them. In the two-vessel system, a differential pressure valve between the two vessels that stages the pressure drop from the wellhead accomplishes this. The energy to maintain this staged pressure is supplied by the multiplex pump through the downhole pump.
The flow path through a cyclone cleaner is shown in Fig. 14.21. Fluid enters the top of the cone tangentially through the feed nozzle and spirals downward toward the apex of the cone. Conservation of angular momentum dictates that the rotational speed of the fluid increases as the radius of curvature decreases, and it is the high rotational speed that cleans the fluid by centrifugal force. The clean fluid, called the overflow, spirals back upward through the vortex core to the vortex finder, while the dirty fluid exits downward at the apex through the underflow nozzle. The cones are usually constructed of cast iron with an elastomer interior. Different feed-nozzle and vortex-finder sizes and shapes are available to alter the performance characteristics of the cyclone. Different sizes of cyclones are available, with the smaller sizes having lower flow-rate capacities but somewhat higher cleaning efficiencies.
Maintaining the proper flow through the cyclone to ensure good cleaning depends on correctly adjusting the pressures at the feed nozzle, overflow, and underflow. At the design flow rates, a 30- to 50-psi drop normally occurs from the feed nozzle to the overflow. In the single-vessel system, a charged pump supplies the pressure, while in a dual-vessel system, the pressure is supplied by a higher backpressure on the returns from the well. Because of the centrifugal head, the cyclone overflow pressure is generally 5 to 15 psi higher than the underflow pressure. An underflow restrictor is commonly used to adjust the amount of underflow to between 5 and 10% of the overflow. This ensures good cleaning without circulation of excessive fluid volumes. It should be noted that the volume flow rates through a cyclone vary inversely with the specific gravity of the fluid, and that within the range of normal power fluids, increased viscosity leads to increased flow rates. The viscosity that suppresses the internal vortex action causes this latter effect. Therefore, proper cyclone sizing to match the charge and multiplex pump characteristics must be done carefully and with detailed knowledge of the fluid to be processed. The manufacturers of the packaged systems supply appropriate cyclones for the installation, but it should be noted that moving the portable unit to another well might require resizing of the cyclone system.
The routing of the dirty underflow varies with different systems, and may be an adjustable option in some systems. Two basic choices are available: return of underflow to the vessel or routing of the underflow to the flowline. In a dual-vessel system, the underflow must be returned to the flowline downstream of the backpressure valve to provide sufficient pressure differential to ensure underflow. Discharging the solids to the flowline is attractive because they are disposed of immediately and are excluded from possible entry into the power fluid. Under some conditions, however, continuous operation may not be possible. If the net well production is less than the underflow from the cyclone for any length of time, the level of fluid in the vessel will drop, and over an extended period of time, this can result in a shutdown of the system. Shutting off the cyclone underflow during these periods stops the loss of fluid, but apex plugging occurs during the shutoff period. Returning the underflow to the vessel eliminates the problem of running the vessel dry but does potentially reintroduce solids into the power fluid. In single-vessel units, the underflow is generally plumbed back to the vessel in a baffled section adjacent to the flowline outlet. This provides for the maximum conservation but requires a differential pressure valve, between the cyclone underflow and the vessel, which is normally set at about 20 psi to ensure a positive pressure to the underflow fluid.
As mentioned previously, the vessel pressure is held above the flowline pressure to ensure flow into the flowline and a backpressure control valve is sometimes used for this purpose. This keeps the vessel pressure, which is backpressure on the well, at a minimum for any flowline pressure that may occur during normal field operation. When water is the power fluid, "riding" the flowline in this manner is acceptable. However, when oil is the power fluid, changing vessel pressure causes flashing of gas in the power oil and adversely affects the multiplex suction. When oil is used as power fluid, it is recommended that a pressure control valve be used to keep the vessel at a steady pressure some 10 to 15 psi above the highest expected flowline pressure.
Although, the single-vessel system was developed for applications involving widely spaced wells, two or three well installations have been successfully operated from a single-well system. This installation is very attractive on offshore platforms. With a large number of highly deviated wells, offshore production is well suited to hydraulic pumping with free pumps, but the extra fluid treating facilities with an open power-fluid system is a drawback when severe weight and space limitations exist. The closed power-fluid system answers this problem, but the extra tubulars in deviated holes create their own set of problems and expense. Furthermore, the use of jet pumps, which is quite attractive offshore, is not possible with the closed power-fluid system. For safety and environmental reasons, water is almost always the power fluid of choice offshore. A large single-well system can receive the returns from all the wells and separate the power water necessary for reinjection to power downhole units. Full 100% separation of the oil from the power water is not necessary, and, in fact, some minor oil carryover will contribute to the power-fluid lubricity. The platform separation facilities then have to handle only the actual production from the wells. A compact bank of cyclone cleaners completes the power-fluid separation and cleaning unit.
In summary, the hydraulic system normally is used in areas where other types of artificial lift have failed or, because of well conditions, have been eliminated because of their shortcomings. Hydraulic pumping systems have been labeled expensive, where, in truth, the use of other artificial lift methods may not be feasible. These include, but are not limited to, the following:
- Using hydraulic free pumps in remote areas where the rig costs are unusually high or the availability of workover rigs is limited.
- Crooked or deviated wells.
- Use of hydraulic systems in relatively deep, hot, high-volume wells. (Note: Hydraulic pumps can go through tubing with as much as a 24° buildup per 100 ft.)
- The use of jet pumps in sandy corrosive wells.
- The use of reciprocating pumps in deep wells with low bottomhole producing pressure.
- Wells with rapidly changing producing volumes.
- The use of jet pumping systems in wells producing with gas/liquid ratios less than 750:1 but producing under a packer where free gas must be pumped.
- Using hydraulic free pumps in wells with high-paraffin contents.
- Using hydraulic OPF systems in low-API-gravity wells.
Jet Pumping System Design ExampleThe following is an example of a design for a well using a jet pumping system. The design data must be carefully collected and is shown in Table 14.9. Because there are numerous possible combinations, and a design typically requires many iterations, current design methods utilize computer software programs.
A jet pumping system was chosen because of the remote location, the advantage of the free-pump system to reduce pump pulling costs, and the advantages and flexibility of a central system to produce several wells drilled in the same field. There are no gas-sales lines, and the produced gas is used to provide the necessary energy to drive the prime movers. The wells are 5,400 ft in depth and have a static reservoir pressure of 2,050 psia. The jet hydraulic pumping system has been operating successfully for 5 years with low operating expenses.
One well was producing only 150 B/D, and a pressure buildup survey and production test indicated a skin of 50. Following a successful reperforating and stimulation treatment, the well is capable of producing significantly higher rates. By running the original jet combination and matching the power fluid, injection pressure, and total production, a new pump intake was calculated, and a new IPR curve was determined.
A design was made to find what could be produced with the existing horsepower and also what might be achieved if excess horsepower from a second well was used. A throat and nozzle (10B) with an annulus of 0.0503 was determined to be a good fit for both cases. See Table 14.6. The selected jet has an ability to produce 1,063 B/D using 1,720 B/D of power fluid at 2,500 psi injection or 81 hp. See Table 14.9. If the power-fluid injection pressure is increased to 3,000 psi, the power-fluid volume is increased to 1,896 B/D, and the pump intake pressure is reduced to 850 psig, then 1,200 B/D of production is feasible, which will take 108 hp.
The predicted performance of the jet pump system for this well is shown in Fig. 14.22. Line 1 on the graph represents 2,500 psi injection and 81 hp. Line 2 represents 3,000 psi and 108 hp. If pressure is increased to 3,500 psi, the pump will go into cavitation, and damage might occur to the jet nozzle throat.
Design Example for a Reciprocating Hydraulic Pump SystemCurrently a 12,000-ft well is equipped with a sucker rod beam pumping system with the pump set at only 9,000 ft. The design data, plus the well completion and pump installation data summary and a pump performance summary, are shown in Table 14.10. The well is deviated with a severe dogleg at 9,100 ft and produces only 100 B/D with a pump intake pressure (PIP) of about 1,000 psi. Workover rig cost is high, and a free-pump installation is desirable to reduce maintenance costs. Furthermore, a production increase is essential for this remotely located well. A review of the IPR data shown in Fig. 14.23 indicates that production can easily be increased from 100 B/D to 350 B/D, if the well can be pumped with a Pwf of 500 psi without significant gas interference. Pressure maintenance operations have begun in the field, and further decrease in the reservoir pressure is not expected. An economic analysis indicates a payout from changing to the hydraulic system in less than 3 years.
The 5½-in. casing has a significant effect on the proposed design. Considering the casing size, depth, production requirements, and reservoir conditions, a casing free-pump system was selected. Power oil is pumped down the tubing and returned up the casing-tubing annulus with the oil, water, and gas production. The 2 7/8-in. [2.441-in. inside diameter (ID)] N-80 tubing now in the well has ample tension, burst, and collapse strengths and will be used. The pump is set at the lowest possible depth (12,000 ft) in order to achieve an operating pressure of 500 psi at the perforations. At design conditions, a pump displacement of about 580 B/D is required to produce the oil and water liquids, plus the free gas. In order to decrease the number of pump failures, the strokes per minute are limited to 33.4. Pump model 252017 was chosen to stay within this range. See Table 14.1.
The selected pump is designed to run at 46.3% of rated speed, requiring a power-fluid volume of 741 B/D and an injection pressure of 3,211.8 psi. Horsepower required for this well is 44.9 hp, and a 60-hp system is selected to provide more flexibility and compensate for wear and possible higher gas volumes.
|An||=||cross-sectional area of nozzle, in.2|
|As||=||cross-sectional area of annulus between throat and jet, in.2|
|At||=||cross-sectional area of throat, in.2|
|D||=||pump setting depth, ft|
|FgL||=||gas/liquid ratio, scf/bbl|
|FmfD||=||dimensionless mass flow ratio|
|gd||=||gradient of return fluid, psi/ft|
|gn||=||gradient of power fluid, psi/ft|
|Pfd||=||friction in discharge tubing, psi|
|Pfpt||=||friction in power tubing, psi|
|Pn||=||useful power fluid pressure at nozzle, psi|
|Pso||=||surface operating pressure, psi|
|Ppd||=||pump discharge pressure, psi|
|Pps||=||pump suction pressure, psi|
|Pwf||=||bottomhole flowing pressure psi|
|Pwh||=||flowline pressure at wellhead, psi|
|qn||=||nozzle flow rate, B/D|
|qs||=||production (suction) fluid rate, B/D|
Figs. 14.1 through14.11 and14.13 through14.21 are reprinted with permission from Weatherford. Copyright © 2004 by Weatherford U.S., L.P. All rights reserved. Weatherford U.S., L.P. disclaims all responsibility for the consequences of any errors or omissions in the materials.
SI Metric Conversion Factors
|°API||141.5/(131.5 + °API)||=||g/cm3|
|bbl||×||1.589 873||E – 01||=||m3|
|ft||×||3.048*||E – 01||=||m|
|hp||×||7.460 43||E – 01||=||kW|
|in.||×||2.54*||E + 00||=||cm|
|in.2||×||6.451 6*||E + 00||=||cm2|
|lbm||×||4.535 924||E – 01||=||kg|
|psi||×||6.894 757||E + 00||=||kPa|
Conversion factor is exact.