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Difference between revisions of "Downhole hydraulic pump types"
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[[Category:3.1.3 Hydraulic and jet pumps]] [[Category:YR]]
[[Category:3.1.3 Hydraulic and jet pumps]] [[Category:YR]]
Latest revision as of 12:40, 4 February 2016
There are two types of pumps used in hydraulic pumping for artificial lift purposes. These are reciprocating pumps and jet pumps. This page discusses each pump type along with a few operational considerations.
The 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. 1. Fig. 2 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. 3.
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 1 through 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 pumps are a type of downhole pump that can be adapted to fit interchangeably into the bottomhole assemblies (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 Brown, Petrie et al.
An example of the simplest downhole jet free-pump completion, the single-seal style, is shown in Fig. 4. 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. 5. 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 bottomhole pressures (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.
Intuitively, 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. 1 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 electrical submersible pumps (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 sizes
Each manufacturer has different sizes and combinations of nozzles and throats. Some manufacturers 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), while others do not. 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 most commonly used area ratios between the nozzle and throat 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 bottomhole pressures (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.
Sizing a jet pump for the application
The 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. 1.
Fig. 2 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.
Deciding which pump type to use
When should a jet be used, and when should a positive-displacement hydraulic pump be used? One possible answer is to use jet pumps if the flowing (pumping) bottomhole pressure (BHP) is large enough because the pressure drawdown capability for the jet system is inferior to that of the reciprocating pump. Other factors enter in as well as those mentioned previously. Jet pumps typically have low pump-repair costs but have high energy-consumption expenses because of low pump efficiencies, usually less than 35%. However, for both systems, a higher pump-failure rate can be very acceptable if a free system is present and the pumps can be retrieved quickly (less than 30 minutes typically) without pulling the tubing.
- Thompson, J. 1852. 1852 Report British Association.
- Gosline, J.E. and O’Brien, M.P. 1933. The Water Jet Pump. U. of California Publication in Eng.
- Gosline, J.E. and O’Brien, M.P. 1933. Application of the Jet Pump to Oil Well Pumping. U. of California Publication in Eng.
- Angier, J.D. and Crocker, F. 1864. Improvement in Ejectors for Oil Wells. US Patent No. 44,587.
- Jacuzzi, R. 1930. Pumping System. US Patent No. 1,758,400.
- McMahon, W.F. 1930. Oil Well Pump. US Patent No. 1,779,483.
- Nelson, C.C. 1975. The Jet Free Pump—Proper Application Through Computer Calculated Operating Charts. Paper presented at the 1975 Southwestern Petroleum Short Course, Texas Tech. U., Lubbock, Texas, 17–18 April.
- Brown, K.E. 1982. Overview of Artificial Lift Systems. J Pet Technol 34 (10): 2384–2396. SPE-9979-PA. http://dx.doi.org/10.2118/9979-PA
- Clark, K.M. 1980. Hydraulic Lift Systems for Low-Pressure Wells. Pet. Eng. Intl.
- Bleakley, W.B. 1978. Design Considerations in Choosing a Hydraulic Pumping System Surface Equipment for Hydraulic Pumping Systems. Pet. Eng. Intl. (July/August).
- Petrie, H., Wilson, P., and Smart, E.E. 1983. The Theory, Hardware, and Application of the Current Generation of Oil Well Jet Pumps. Paper presented at the 1983 Southwestern Petroleum Short Course, Texas Tech. U., Lubbock, Texas, 27–28 April.
- Petrie, H., Wilson, P., and Smart, E.E. 1983. Jet Pumping Oil Wells. World Oil (November–December 1983; January 1984).
- Kempton, E.A. 1980. Jet Pump Dewatering, What it is and How it Works. World Oil (November).
- Cunningham, R.G. and Brown, F.B. 1970. Oil Jet Pump Cavitation. Paper presented at the 1970 ASME Cavitation Forum, Joint ASME Fluids Engineering, Heat Transfer, and Lubrication Conference, Detroit, Michigan, 24–27 May.
Noteworthy papers in OnePetro
Use this section to list papers in OnePetro that a reader who wants to learn more should definitely read
Bradley, H. B., & Gipson, F. W. (1992). Petroleum engineering handbook. Richardson, TX, U.S.A: Society of Petroleum Engineers. WorldCat
Frick, T. C., & Taylor, R. W. (1962). Petroleum production handbook. Dallas, Tex: Society of Petroleum Engineers of AIME. WorldCat
Pugh, Toby. (2014). Overview of Hydraulic Pumping. Weatherford. iBook.
Use this section to provide links to relevant material on websites other than PetroWiki and OnePetro