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This page discusses the specific artificial-lift technique known as beam pumping, or the sucker-rod lift method. Many books, technical articles, and industry standards have been published on the sucker-rod lift method and related technology.artificial lift system. The Gipson and Swaim “Beam Pump Design Chain” is used as a foundation and built upon using relevant, published technology.

Beam pumping, or the sucker-rod lift method, is the oldest and most widely used type of artificial lift for most wells. A sucker-rod pumping system is made up of several components, some of which operate aboveground and other parts of which operate underground, down in the well. The surface-pumping unit, which drives the underground pump, consists of a prime mover (usually an electric motor) and, normally, a beam fixed to a pivotal post. The post is called a Sampson post, and the beam is normally called a walking beam. Figs. 1 and 2 present a detailed schematics of a typical beam-pump installation.

This system allows the beam to rock back and forth, moving the downhole components up and down in the process. The entire surface system is run by a prime mover, V-belt drives, and a gearbox with a crank mechanism on it. When this type of system is used, it is usually called a beam-pump installation. However, other types of surface-pumping units can be used, including hydraulically actuated units (with and without some type of counterbalancing system), or even tall-tower systems that use a chain or belt to allow long strokes and slow pumping speeds. The more-generic name of sucker-rod lift, or sucker-rod pumping, should be used to refer to all types of reciprocating rod-lift methods.

Linked rods attached to an underground pump are connected to the surface unit. The linked rods are normally called sucker rods and are usually long steel rods, from 5/8 to more than 1 or 1 1/4 in. in diameter. The steel rods are normally screwed together in 25- or 30-ft lengths; however, rods could be welded into one piece that would become a continuous length from the surface to the downhole pump. The steel sucker rods typically fit inside the tubing and are stroked up and down by the surface-pumping unit. This activates the downhole, positive-displacement pump at the bottom of the well. Each time the rods and pumps are stroked, a volume of produced fluid is lifted through the sucker-rod tubing annulus and discharged at the surface.

Sucker-rod pumping systems should be considered for new, lower volume stripper wells, because they have proved to be cost effective over time. Operating personnel usually are familiar with these mechanically simple systems and can operate them efficiently. Inexperienced personnel also can operate rod pumps more effectively than other types of artificial lift. Most of these systems have a high salvage value.

Understanding the makeup of the producing reservoir, its pressure, and the changes that occur in it are important to attain maximum production. Because reservoir conditions change as fluids are produced, ongoing measurement of the reservoir conditions is necessary. The main considerations in measuring and understanding the reservoir are the types and volumes of reservoir fluids being produced, their pressures in both the reservoir and at the wellbore or pump intake, and the effects these fluids have as they pass through the producing system.

A variety of well tests and measurements may be used to determine production rates for oil-, gas-, and water-supply wells and to observe the status of the reservoir. Each test reveals certain information about the well and the reservoir being tested. The main reservoir considerations are determining bottomhole pressure and the inflow relationship of the fluids with changing reservoir and pump-intake pressure.

Bottomhole pressure measuring equipment (pressure bombs) makes it possible to determine reservoir and tubing intake pressures within the desired range of accuracy. When this test is conducted at scheduled intervals, valuable information about the decline or depletion of the reservoir from which the well is producing can be obtained. However, it is difficult to obtain either bottomhole reservoir or operating pressures while the rod-pump system is installed and operating.

The key to accurate bottomhole pressure determination in any pumping well is the ability to predict the gradient of the fluid in the casing/tubing annulus. In 1955, W.E. Gilbert (Unpublished internal report: “Curve Annulus Gradient Correction for Gas Bubbling Through Static Liquid Column,” Shell Oil Co.) developed an iterative calculation procedure on the effect of gas bubbling up a static fluid column. This can be used in a trial-and-error method to determine a gradient correction factor (F) to determine the pressure at the desired depth in the presence of gas production. If the term Q /(aP)0.4 is greater than 0.25, this method should be used with caution because this is an indication that liquid flow up the annulus may occur. Also, the crude pressure/volume/temperature (PVT) characteristics alter the results. The Gilbert curve and a calculation example are presented in "The Beam Pump Design Chain."

Knowing the reservoir and pump intake pressures during static and operating conditions will allow a determination of the well"s production capacity. This is required to optimize the artificial-lift equipment and properly size the equipment that is installed. The well productivity under varying production conditions must then be known.

One of the most critical decisions in an artificial-lift system is the selection and design of equipment appropriate for the volume of fluid the reservoir produces. Reservoir inflow performance detaisl the productivity index and IPR of fluids with changes in reservoir pressure. Because most fluid produced by an artificial lift method is not single phase, it is not in a steady-state condition. Also, because most pumping operations occur after the fluid is below the bubblepoint pressure, the IPR method is usually considered. This technique takes into account various fluid phases and flow rates. It was originally devised by Vogel

Producing rates can be estimated within the desired range of accuracy using the IPR technique with two stabilized producing rates and corresponding stabilized producing pressures. This makes it possible to use the IPR without needing to shut in the well and lose production to obtain shut-in information. Obtaining a bottomhole pressure equal to 10% of the shut-in reservoir pressure is recommended for determining maximum production rates for sucker-rod lifted wells. At this pressure, the maximum well productivity will be 97% of the well"s theoretical maximum production rate. However, the maximum lift-design rate should, in most cases, be slightly higher to permit some downtime and decreased pump efficiency.

In any artificial lift system, the volume of gas produced should be considered in designing the system and in analyzing the operation after the system has been installed. A complete analysis requires knowing the volume of gas in solution, the volume of free gas, the formation volume factors, and whether gas is produced through the pump or is vented. If PVT analyses of reservoir fluids are available, they are the most accurate and easiest to use as a source of solution gas/oil ratio (GOR), formation volume factors, etc. The next best source is an analysis from a nearby similar reservoir.

When pumping through tubing in the absence of a production packer, free gas, which breaks out of the oil, should be vented up from the casing/tubing annulus. However, when it is necessary to produce from beneath a production packer, a vent string can be installed. The possibility of needing a vent string should be considered when planning casing sizes for a new well.

Both the size of the vent string and the location of its bottom, with respect to the location of the pump intake and producing perforations, will influence the string"s effectiveness in removing free gas. The string"s diameter should be designed to allow the production of the anticipated free-gas volume with a pressure drop no greater than the desired producing bottomhole pressure minus the surface backpressure. If the required pressure drop is greater than this, a portion of the free gas will have to go through the pump. Fig. 2 is an indication of the effect of vent-string size on the pressure drop through it. Care should be taken if small-diameter tubing is used, because it may not allow all the gas to flow up the vent or may simply load up and prevent most gas flow.

Gas that remains in solution when the liquid enters the pump increases the volume of total fluid through the pump compared to the liquid measured at the surface by the formation volume factor at pump-intake conditions. The gas also decreases the density of the fluid and, thus, the head or pressure to be pumped against in the tubing. Free gas that enters the pump must be compressed to a pressure equivalent to the head required to lift the fluid. This free gas will reduce the volume of both the produced liquid that enters the pump and the liquid measured at the surface. Any time the pump does not compress the free gas to a pressure greater than that exerted on the pump by the fluid column in the producing string, production ceases and the pump is said to be "gas locked." This condition can exist in both plunger and centrifugal pumps.

Intake pressure is the pressure in the annulus opposite the point at which the fluid enters the pump. If the pump intake pressure is increased by increasing the pump submergence, the free gas volume decreases because the fluid retains more gas in solution. Reducing the pressure drop in the pump-suction piping also reduces the free gas to be produced. The pump intake should not be deeper than is necessary to maintain the desired intake pressure. A pump intake that is too deep results in unnecessary investment and increased operating costs.

Fig. 3 is a graph of the liquid produced as a percent of the displacement of a plunger pump plotted against the pump intake pressure for a typical reservoir.

Fig. 3—Example of liquid produced as a percentage of plunger-pump displacement for various pump-intake pressures and the effects of gas on efficiency.

Gas bubbles entrained in the produced liquid(s) tend to rise because of the difference in the liquid and gas densities. The rate of bubble rise depends on the size of the bubbles and the physical properties of the fluid. The size of the bubbles increases as the pressure decreases. At low pump-intake pressures, the rate of gas-bubble rise in low-viscosity fluids will approximate 0.5 ft/sec, assuming a 400-μm bubble rise in water. The increase in bubble size and rate of rise as the pressure decreases causes the reversal in curves B–D and B–E in Fig. 3.

Downhole gas separators are used in gassy wells to increase the volume of free gas removed from the liquids before reaching the pump. However, they are not 100% effective in separating the gas. In sucker-rod-pumped wells, these separators are normally called "gas anchors." Gas anchors are usually designed and built in the field; Fig. 4 contains schematic drawings of six common types. The most commonly used are the "natural" gas anchor (A) and the "poor boy" gas anchor (C). Typically, there are two major components for these gas anchor assemblies, the mud anchor run on the bottom of the tubing string and the dip tube or strainer nipple run on the bottom of the pump.

The largest downhole gravity separator is normally the casing/tubing annulus. This area provides a maximum down passage for liquid and up-flow area for gas. This allows the oil (and water) to move relatively slowly, typically, downward from the perforations to the pump, and permits the gas to separate and flow upward. For this reason, a natural gas anchor should be used whenever practical because it takes advantage of the entire casing internal cross-sectional area. This type of separator typically should be placed approximately 15 ft below the lowest most-active well perforations. However, if there is insufficient distance in the well to place the pump intake below the perforations, then the pump intake should be placed approximately 15 ft above the top-most perforation and a poor boy separator should be properly designed and installed.

There are limitations on how much gas can be handled by the downhole separator. If more gas is produced than can be handled by the separator, the gas will not separate completely. The downhole pump must then handle the excess gas. If the wells exceed these theoretical gas rates, then:

The situation worsens if excessive gas enters the pump and there is insufficient compression ratio to pump all the fluids, resulting in a gas-locked pump. When this occurs, operating costs for this well increase dramatically because when there is no production, there is no revenue. However, a properly designed and spaced pump should not gas lock if the well is not pumped off.

It is often recommended that the outside diameter (OD) of the gas anchors" steel mud anchor be less than the ID of the largest overshot or wash pipe that can be run in the well casing. This limits the gas-anchor separation capacity that can be secured in wells with small casings. Reinforced plastic mud anchors that can be drilled up, or steel designs that can be recovered with spears, should be considered when mud anchor OD must approach casing-drift diameter. This design would then be considered the "modified poor boy." Agreement should be obtained from the field before installation to ensure acceptance of the possible problems when trying to pull this type of installation.

In 1954, an in-depth study of the complex aspects associated with sucker rod pump design was started. Through this effort, Sucker Rod Pumping Research, Incorporated, a non-profit organization was created. The services of the Midwest Research Institute at Kansas City were retained to perform the work necessary to achieve the objectives of the organization. Midwest Research Institute published its report in 1968, which was then used to create the industry standard API RP 11L. Gipson and Swaim did an excellent job of summarizing a sucker-rod lift-system design in The Beam Pump Design ChainRP 11L approach. API RP 11L is superseded by API TR 11L. This recommended practice should be consulted for continued discussion of this equipment, along with a review of a sample problem and a recommended solution. Prior to this, Gibbs (1963) introduced a solution for wave equation that simulates force wave propagation through sucker rod string. The approach has been enormously updated since then by multiple authors to consider further details of the physics of the phenomena and to enhance capturing the effect of fluid properties. . The approach has become the base for multiple commercial beam pump design software.

In summary, use the design procedure presented in API TR 11L or a suitable wave equation. Several commercial wave-equation computer programs are available that many operators have successfully used. In the following, the beam pump design procedure based on API TR 11L is introduced. Further details are found in Takacs (2015).

Due to the elasticity of the rod, the rod string might strength or contract through the pumping cycle. This results in a downhole stroke length at the plunger "Sp" that slightly differs from the design stroke length S. This difference results in an actual flow "qa" that is different from the design flow rate "q". Based on API TR 11L, the rod stretch is predicted. "qa" is then calculated and is compared to the desired "q". The optimum "q" can then be reached with an iterative procedure. The procedure or this calculation stats with determining the theoretical flow rate "q" from the pump speed "N", surface stroke length "S" , and plunger size "dp" as follows,

A primary selection of rod string design is required. Firstly, the total length of rod string approximately equals the pump setting depth L in non-literal wells. Moreover, the configuration of rod string diameters is determined from a standard set of configurations provided in API RP 11L. The standard provides a table of the characteristics of the tapered rod string. Figure 1 shows a portion of table 4.1 provided for rod string configuration and properties. The figure is a snapshot of the digitized table provided by the Petroleum Extension (PETEX®) of The University of Texas at Austin, which can be found here under the title “Beam Lift System Design Calculators.” In Figure 1, the last six columns are API sizes of rod diameters. The table numbers represent the percentage of each size in the making of the rod string. the first column is a list of configuration identifiers. Based on dp determined in the previous step, a rod configuration is selected. The other information provided by the table for each rod configuration is

The dimensionless number Sp/S is defined in API TR 11L as a function of two other dimensionless numbers, namely N/No" and F0/Skr. N/No" condenses the effect of pumping speed and natural frequency in the tapered rod strings. The natural frequency of non-tapered rod string No is defined by Griffin (1968) as the number of strokes that propagates through the rod string at four times the velocity of sound during the unit time. Therefore, it takes the frequency unit, namely, strokes per unit time. It is mathematically written as,

where kt is the Spring Constant of the unanchored tubing and represents the load required to stretch the unanchored portion of the tubing, between the anchor and the pump, unit length. Similar to Eq. (6), kt is defined as

From Sp/S = Sp/S x S, "qa" is calculated using Eq. (1). If not acceptable, "N", "S" , or "dp" are changed and and an iterative procedure is started from step 1. Increasing "N" to compensate for stroke length loss does not come free of expense. The more "N" is increased, the shorter the rod string and pump fatigue life will be. Moreover, increasing "dp" results in a shorter Sp due to inertia effects. Therefore, an optimum selection of these parameters is needed.

Throughout the pump cycle, the polished rod exhibits varying loads that swing between two extremities, namely, the Maximum Polished Rod Load PPRL and the Minimum Polished Rod Load MPRL. PPRL and MPRL are found as follows,

Gibbs, S. G. 1963. Predicting the Behavior of Sucker-Rod Pumping Systems. Journal of Petroleum Technology, 15(7), 769-778. https://doi.org/10.2118/588-PA.

Griffin, F. D. 1968. Electric Analog Study of Sucker-Rod Pumping Systems. Paper presented at the Drilling and Production Practice, New York, New York.

Gipson, F.W. and Swaim, H.W. 1988. The Beam Pumping Design Chain. Paper presented at the 1988 Southwestern Petroleum Short Course, Lubbock, Texas, 23–25 April.

Hein Jr., N.W. 1996. Beam-Pumping Operations: Problem Solving and Technology Advancements. J Pet Technol 48 (4): 330-336. SPE-36163-MS. http://dx.doi.org/10.2118/36163-MS

McCoy, J.N., Podio, A.L., and Becker, D. 1992. Pressure Transient Digital Data Acquisition and Analysis From Acoustic Echometric Surveys in Pumping Wells. Presented at the Permian Basin Oil and Gas Recovery Conference, Midland, Texas, 18-20 March 1992. SPE-23980-MS. http://dx.doi.org/10.2118/23980-MS

Downhole Diagnostic. "Sucker Rod Pumping Wells: Design, Operation, & Optimization." Scribd. http://www.scribd.com/doc/238486620/Sucker-Rod-Pumping-Wells-Design-Operation-Optimization.

Centrifugal pumps are the most commonly used kinetic-energy pump. Centrifugal force pushes the liquid outward from the eye of the impeller where it enters the casing. Differential head can be increased by turning the impeller faster, using a larger impeller, or by increasing the number of impellers. The impeller and the fluid being pumped are isolated from the outside by packing or mechanical seals. Shaft radial and thrust bearings restrict the movement of the shaft and reduce the friction of rotation.

Centrifugal pumps are designed with respect to the number of suctions (single or double), number of impellers (single, double, or multistage), output, and impellers (type, number of vanes, etc.). Most impellers are arranged from one side only and are called single-suction design. High-flow models use impellers that accept suction from both sides and are called double-suction design.

The efficiency of a centrifugal pump is determined by the impeller. Vanes are designed to meet a given range of flow conditions. Fig. 6.5 illustrates the basic types of impellers.

Open Impellers. Vanes are attached to the central hub, without any form, sidewall, or shroud, and are mounted directly onto a shaft. Open impellers are structurally weak and require higher NPSHR values. They are typically used in small-diameter, inexpensive pumps and pumps handling suspended solids. They are more sensitive to wear than closed impellers, thus their efficiency deteriorates rapidly in erosive service.

Partially Open (Semiclosed) Impellers. This type of impeller incorporates a back wall (shroud) that serves to stiffen the vanes and adds mechanical strength. They are used in medium-diameter pumps and with liquids containing small amounts of suspended solids. They offer higher efficiencies and lower NPSHR than open impellers. It is important that a small clearance or gap exists between the impeller vanes and the housing. If the clearance is too large, slippage and recirculation will occur, which in turn results in reduced efficiency and positive heat buildup.

Closed Impellers. The closed impeller has both a back and front wall for maximum strength. They are used in large pumps with high efficiencies and low NPSHR. They can operate in suspended-solids service without clogging but will exhibit high wear rates. The closed-impeller type is the most widely used type of impeller for centrifugal pumps handling clear liquids. They rely on close-clearance wear rings on the impeller and on the pump housing. The wear rings separate the inlet pressure from the pressure within the pump, reduce axial loads, and help maintain pump efficiency.

Single-Stage Pumps. The single-stage centrifugal pump, consisting of one impeller, is the most widely used in production operations. They are used in pumping services of low-to-moderate TDHs. The TDH is a function of the impeller’s top speed, normally not higher than 700 ft/min. Single-stage pumps can be either single or double suction. The single-stage pump design is widely accepted and has proved to be highly reliable. However, they have higher unbalanced thrust and radial forces at off-design flow rates than multistage designs and have limited TDH capabilities.

Multistage Pumps. The multistage centrifugal pump consists of two or more impellers. They are used in pumping services of moderate-to-high TDHs. Each stage is essentially a separate pump. All the stages are within the same housing and installed on the same shaft. Eight or more stages can be installed on a single horizontal shaft. There is no limit to the number of stages that can be installed on a vertical shaft. Each stage increases the head by approximately the same amount. Multistage pumps can be either single or double suction on the first impeller.

A single-suction, enclosed or semienclosed impeller is inherently subject to continual end thrust. The thrust is directed axially toward the suction because of the low pressures that exist in the impeller eye during pump operation. This thrust is handled with a thrust bearing. The larger the TDH and the larger the impeller-eye diameter, the larger the thrust. Excessive thrust results in bearing and seal damage.

Thrust can be reduced by designing a single-stage impeller for a double suction. In multistage pumps, thrust can be reduced by facing half the impellers in one direction and half in the other. Balancing holes can be used in single-suction, single-stage pumps. The impeller is cored at the rear shroud to allow high-pressure liquid to flow back to the impeller eye.

As the fluid leaves the top of the rotating impeller, it exerts an equal and opposite force on the impeller, shaft, and radial bearings. At the best-efficiency point (BEP), the sum of all radial forces nearly cancels each other out. At capacities below or above the BEP, forces do not cancel out completely because the flow is no longer uniform around the periphery on the impeller. Radial forces can be significant. Heavy-duty radial bearings may be required in lieu of the manufacturer’s standard when pump operation departs significantly from the BEP.

Pump specific speed is the speed in revolutions per minute required to produce a flow of 1 gal/min with a TDH of 1 ft, with an impeller similar to the one under consideration but reduced in size. The pump specific speed links the three main components of centrifugal-pump performance characteristics into a single term. It is used to compare two centrifugal pumps that are geometrically similar. Pump specific speed can be calculated from

The pump specific speed is always calculated at the pump’s point of maximum efficiency. The number is used to characterize a pump’s performance as a function of its flowing parameters. Normally, it is desirable to select the impeller with the highest specific speed (smallest diameter). This may be offset by the higher operating cost associated with higher speeds and greater susceptibility to cavitation damage.

Impellers With Low Specific Speeds (500 to 4,000). Radial-flow impellers typically have low specific speeds. Radial-flow impellers are narrow and relatively large in diameter and are designed for high TDHs and low flow capacity. The pumped fluid undergoes a 90° turn from inlet to outlet of the impeller.

Impellers With Median Specific Speeds (4,000 to 10,000). Mixed-flow impellers typically have medium specific speeds and are wider and smaller in diameter than radial-flow impellers. They exhibit medium TDH and medium flow capability. They are typically used in vertical multistage pumps and downhole electrical submersible pumps, which require small diameters.

Impellers With High Specific Speeds (10,000 to 16,000). Axial-flow impellers typically have high specific speeds. In these impellers, the liquid flow direction remains parallel to the axis of the pump shaft. Axial-flow impellers are used for high flow and low TDH applications. They are most commonly used for water irrigation, flood control, pumped storage power-generation projects, and as ship impellers.

When a pump manufacturer develops a new pump, the new pump is tested for performance under controlled conditions. The results are plotted to show flow rate vs. head, efficiency, and power consumption. These graphs are known as performance curves. Under similar operating conditions, an installed pump is expected to demonstrate the same performance characteristics as shown on the performance curves. If it does not, this indicates that something is wrong with the system and/or pump. Comparison of actual pump performance with rated performance curves can help determine pump malfunction.

Curve Performance. The impeller shape and speed is the primary determinant of pump performance. Fig. 6.6 illustrates a generalized centrifugal-pump curve. Head, NPSHR, efficiency, horsepower, and brake-horsepower (BHP) requirements vary with flow rate. The TDH is greatest at zero capacity (shutoff head) and then falls off with increasing flow rates. The horsepower curve starts out at some small value at zero flow, increases moderately up to a maximum point, and then tapers off slightly. The pump efficiency curve starts out at zero, increases rapidly as flow increases, levels off at the BEP, and decreases thereafter. The NPSHR is a finite value at zero flow and increases as the square of the increase in flow rate.

Curve Parameters. It is best to operate the pump at the BEP, but this is not normally feasible. Alternatively, the pump should operate only in the area of the curve closest to the BEP and only in the moderately sloping portion of the head curve. Operating in the flat or steeply sloping portions of the curve results in wasted energy and flow control instability. Pumps that run at or near BEP run smoother and have better run lives. Any time the actual flow drops to less than 50% of the BEP flow, it is wise to consult the manufacturer because shaft deflections may increase dramatically (especially with single-stage overhung-design pumps), which could lead to higher maintenance costs and to failures.

Pumps in Parallel. Fig. 6.7 illustrates the shape the TDH-vs.-capacity curve assumes when identical pumps are operated in parallel and series. Parallel operation occurs where multiple pumps are piped to the same suction and discharge lines. The combined flow rate is the total of the individual pump flows at the TDH. In most cases, the head capacity curves of the parallel pumps are the same, or nearly so. It is not necessary for the curves to be the same as long as each pump operating in parallel can put out the desired TDH.

All centrifugal pumps discharging to an elevated or pressurized vessel and all centrifugal pumps operating in parallel should have check valves in the event of a pump shutdown to keep the pump from spinning backwards. (The danger is a sheared shaft on restart attempt.)

Driver size should be selected so that overloading does not occur at any point across the entire pump curve. Flow orifices or meters should be provided in each pump’s discharge line for verification of flow rates. Suction and discharge piping should be arranged as symmetrically as practical so that all pumps have the same NPSHA.

Series Operation. Series operation is used when a single pump cannot develop the total TDH required. It is also used when a low NPSHR is used to feed a larger pump that requires an NPSHR that cannot be provided from an atmospheric tank or vessel operating at its bubblepoint. In series operation, the combined head is the sum of the individual-pump TDHs at the same flow.

The system head curve is a graphical representation of TDH required to be furnished by the pump vs. the flow rate through the piping system. It consists of a constant (static) and an increasing (variable) portion. Fig. 6.8 illustrates an example of a typical system head curve.

It is unusual for a system to require operation at a single fixed flow rate. A pump will deliver only the capacity that corresponds to the intersection of the TDH capacity and system head curves. To vary the capacity, one must change the shape of one or both curves. The head-capacity-curve shape can be changed by altering the pump speed or impeller diameter. The system-head-curve shape can be changed by the use of a backpressure throttling valve (see Sec. 6.3.11).

The effects of operating at significantly reduced capacity may lead to operating at much less than the BEP, higher energy consumption per unit capacity, high bearing loads, temperature rise, and internal circulation. These results can be minimized with the use of a variable-speed driver or with the use of several parallel pumps for the total capacity and sequentially shutting down individual units as demand requires.

Higher bearing loads will exist for any flow that departs from the BEP, especially for single-stage, single-suction pumps. This can be anticipated by specifying certain types of heavy-duty and long-life bearings. If the temperature of the pumped fluid rises and the flow rate through the pump decreases, minimum-flow recirculation can be used (see Sec. 6.3.12). The manufacturer generally provides the minimum continuous required flow rate for any pump selection. Operating between the BEP and minimum required flow rate generally avoids all the problems discussed.

The difference between the TDH developed by the pump and the head required by the system head curve represents lost energy. Because most centrifugal pumps are driven by constant-speed electric motors, throttling is the only practical method of regulating capacity. The backpressure valve imposes a variable amount of loss on the system head curve. Closing the valve increases control losses and causes the system head curve to slope up more steeply to intersect the TDH capacity curve at the desired capacity. Opening the valve decreases the control losses and causes the system head curve to slope downward and intersect the TDH capacity curve at a higher capacity. With the valve completely open, the capacity is governed only by the intersection of the two curves.

The recirculation valve prevents the buildup of excessive amounts of heat within the casing. A minimum-flow recirculation valve should be installed if the pump piping system contains a backpressure valve that could close and result in less than the minimum continuous flow at which the pump can safely operate. A recirculation valve is often used in installations in which the pump piping contains an automatic shutdown discharge valve that could fail in the closed position, or a discharge block valve that can be inadvertently closed. The recirculation valve should be upstream of the first block valve or control valve downstream of the pump. On small pumps, an orifice is usually installed on the recirculation, which continuously recirculates a fixed flow of liquid back to the suction. A control valve costs more but will modulate the recirculation to assure only minimum flow and thus result in less energy loss.

The maximum head that a centrifugal pump can develop is determined by speed, impeller diameter, and number of stages. Thus, to change the head of a pump, one or more of these factors must be changed. Speed can be changed with different gears, belts, or pulleys, or by installing a variable-speed driver. The impeller diameter can be altered for large permanent changes. The number of impellers can be changed by replacing existing impellers with spacers or dummy impellers.

Most motor-driven centrifugal pumps are operated at constant speed. A direct-current or variable-frequency alternating-current motor control can maintain nearly the same pump efficiency over a larger speed range. Variable-speed control makes it possible to eliminate the backpressure throttling requirements to adjust system head.

Fig. 6.9 illustrates the head-capacity-curve relationship of a constant-speed and variable-speed pump. The pump is operating at 100% of its capacity, and the TDH is represented by Point 1 on the graph. If it becomes desirable to reduce the capacity to 80% of the rated capacity, the constant-speed-pump operation will move to Point 3. Point 3 requires 110% of the head and 92% of the BHP required at Point 1, and thus, additional backpressure would be required to force the system curve to intersect the pump curve at this point.

A variable-speed driver could, in effect, find a TDH capacity curve that intersects the system curve at Point 2. Point 2 requires only 70% of the head and 73% of the power required at Point 1. Thus, at 80% capacity, the constant-speed pump would operate at Point 3 and the variable-speed pump at Point 2. The potential energy savings is represented by the difference between 92 and 73% of horsepower, or 19%.

The affinity laws are used to predict what effect speed or impeller-diameter changes have on centrifugal-pump performance. The laws are based on dimensional analysis of rotating machines that shows, for dynamically similar conditions, certain dimensionless parameters remain constant. These relationships apply to all types of centrifugal and axial machines.

Predictions for speed changes are fairly accurate throughout the range of speed changes. However, predictions for diameter changes tend to be accurate for diameter change of only ± 10% because changing the diameter also changes the relationship of the impeller to the pump casing. Thus, for a 10% increase in either diameter or speed, the flow will increase by 10%, TDH by 21%, and the BHP by 33%.

Most centrifugal pumps have a flooded suction. The source is above the pump suction, and atmospheric pressure is sufficient to maintain fluid at the pump inlet at all times. Sometimes the pump must take suction from a source that is below the centerline of the pump. Atmospheric pressure alone will not always keep the suction flooded. Conventional centrifugal pumps are not self-priming. Thus, they are not capable of evacuating vapor from the casing so that fluid from the suction line can replace the vapor. Self-priming pumps are designed so that an adequate fluid volume for repriming is always retained within the pump casing, even if fluid drains back to the source.

A centrifugal pump is a piece of precision machinery that must not be subjected to external strains beyond those it was designed to encounter. It must be installed in the intended position, carefully aligned, and free from piping forces and moments.

Foundations. Generally, foundation design is not critical. Vibration in a centrifugal pump is minimal unless an engine driver is used. As a general rule of thumb, the foundation should be able to handle three times the weight of the pump, driver, and skid assembly. The manufacturer is the best source for determining the required foundation size.

Piping Design. Poor piping design and installation is a common cause of poor centrifugal-pump performance or failure. Poor piping can result in cavitation, performance dropout, impeller failure, bearing and mechanical seal failures, cracked casings, and leaks, spills, and fires.

Fluid-Source Inlet. When the fluid source is above the pump (static head), the source vessel should contain a weir to minimize turbulence, a vortex breaker to eliminate vortexing and vapor entrainment, and a nozzle sized to limit exit velocity to 7 ft/sec or, preferably, less. When the fluid source is below the pump (static lift), the sump, basin, or pit should be designed to provide even velocity distribution in the approach or around the suction inlet and should be sufficiently submerged to prevent vortexing.

Pipe Size and Elimination of Air Pockets. Piping should be at least one nominal pipe size larger than the pump suction flange. Velocities should be less than 2 to 3 ft/sec, and the head loss as a result of friction should be less than 1 ft per 100 ft of equivalent piping length. Suction lines should be short and free of all unnecessary turns. For flooded suctions, piping should be continuously sloping downward to the pump suction so that any vapor pockets can migrate back to the source vessel. For static lifts, the piping should be continuously sloping upward with no air pockets (install gate valves in horizontal position). Where air pockets cannot be avoided, the use of automatic vent valves is recommended.

Upstream Elbow Considerations. When making upstream orientation changes, only long-radius elbows should be used. They should not be connected directly to the pump suction flange, and a minimum of at least two to five pipe diameters of straight pipe should be between the suction flange and the elbow and between successive elbows. This reduces swirl and turbulence before the fluid reaches the pump. Otherwise, separation of the leading edges may occur, with consequent noisy operation and cavitation damage.

Eccentric Reducers. Reducers are required when making a transition from one pipe size to another and in going from the suction-pipe size to the pump flange. Reduction at the pump should be limited to one nominal size change (e.g., 8 to 6 in.). If two or more nominal pipe size reductions are required, it is best to locate any remaining changes several pipe diameters away from the pump inlet. Eccentric reducers should be used, if possible, and should be installed with the flat side up. Concentric reducers should not be used for horizontal suction lines because they could trap vapor that can be pulled into the pump and cause cavitation or vapor lock. Concentric reducers can be used for vertical suction lines and horizontal lines with flooded suction.

Discharge Piping. Minimum Flow Bypass. The minimum-flow bypass (or "recirculation") protects the pump from temperature buildup when the pumping rates are low. They should be designed to handle the pump’s minimum flow capacity at minimum discharge pressure with a line restrictor to adjust flow. Small pumps are usually controlled by an orifice or choke tube. For large pumps in which a continuous bypass would consume excessive power, a control valve actuated (opened) by low flow is used.

Check Valves. Check valves are essential to minimize backflow, which can damage the pump. Selection should take into account the effect of water hammer. Water hammer is the transient change in static line pressure as a result of a sudden change in flow. Items that can start the sudden change in flow include the starting or stopping of a pump or the opening or closing of a check valve.

Slow-closing check valves are acceptable on systems with a single pump and long lengths of pipe. Fast-closing check valves are required with multiple pumps operating in parallel and at high heads. As a general guideline, lift ("swing") check valves are slow unless they are spring loaded. Tilting-disk check valves are fast closing but are more expensive and have a higher pressure drop than swing check valves. When fast-reacting check valves are required, pressure-drop considerations should be secondary.

The normal ESP system configuration is shown in Fig. 13.1. It shows a tubing-hung unit with the downhole components comprising of a multistage centrifugal pump with either an integral intake or separate, bolt-on intake; a seal-chamber section; and a three-phase induction motor, with or without a sensor package. The rest of the system includes a surface control package and a three-phase power cable running downhole to the motor. Because of the ESP’s unique application requirement in deep, relatively small-bore casings, the equipment designer and manufacturer are required to maximize the lift of the pump and the power output of the motor as a function of the diameter and length of the unit. Therefore, the equipment is typically long and slender. The components are manufactured in varying lengths up to approximately 30 ft, and for certain applications, either the pump, seal, or motor can be multiple components connected in series.

Throughout their history, ESP systems have been used to pump a variety of fluids. Normally, the production fluids are crude oil and brine, but they may be called on to handle liquid petroleum products; disposal or injection fluids; and fluids containing free gas, some solids or contaminates, and CO2 and H2S gases or treatment chemicals. ESP systems are also environmentally esthetic because only the surface power control equipment and power cable run from the controller to the wellhead are visible. The controller can be provided in a weatherproof, outdoor version or an indoor version for placement in a building or container. The control equipment can be located within the minimum recommended distance from the wellhead or, if necessary, up to several miles away. API RP11S3 provides the guidelines for the proper installation and handling of an ESP system. Table 13.1, some of which are discussed later in this chapter.

The ESP is a multistage centrifugal type. A cross section of a typical design is shown in Fig. 13.2. The pumps function is to add lift or transfer pressure to the fluid so that it will flow from the wellbore at the desired rate. It accomplishes this by imparting kinetic energy to the fluid by centrifugal force and then converting that to a potential energy in the form of pressure.

In order to optimize the lift and head that can be produced from various casing sizes, pumps are produced in several diameters for application in the most common casing sizes. Table 13.2 lists some common unit diameters, flow ranges, and typical casing sizes in which they fit.

Shaft. The shaft is connected to the seal-chamber section and motor by a spline coupling. It transmits the rotary motion from the motor to the impellers of the pump stage. The shaft and impellers are keyed, and the key transmits the torque load to the impeller. As was mentioned earlier, the diameter of the shaft is minimized as much as possible because of the restrictions placed on the pump outside diameter. Therefore, there are usually several shaft material options available, depending on the maximum horsepower (HP) load and corrosion protection required.

Housing. The housing is the pressure-containing skin for the pump. It holds and aligns all the components of the pump. There are several material options available for different application environments. For additional corrosion protection, there are several coatings that can be applied.

Several different styles of intakes can be selected. They allow for entrance of the fluid into the bottom of the pump and direct it into the first stage. Integral intakes can be threaded directly into the bottom of the housing during the manufacturing assembly process, while others are separate components, which are bolted on to the bottom pump flange.

A standard intake has intake ports that allow fluid to enter the pump. It is used when the fluid is all liquid or has a very low free-gas content. The intake shown in Fig 13.2 would be a standard intake if the reverse-flow screen were omitted.

A reverse-flow intake is used when the free-gas content in the fluid is high enough to cause pump-performance problems. The pump in Fig. 13.2 is shown with a reverse-flow design. The produced fluid with free gas flows up the outside of the reverse-flow intake screen, makes a 180° turn to enter through the perforations or holes at the top of the screen, flows back down to the intake ports and then back up to the first pump stage. These reversals in direction allow for a natural separation of the lighter gases from the liquid. The separated gas travels up the casing annulus and is vented at the wellhead. Another style is shown in the right-hand graphic of Fig. 13.3, which has a longer reversing path than does the intake with the screen.

The next step in handling free gas with an ESP involves downhole mechanical separation devices such as separator intakes. These devices take the fluid that enters its intake ports, impart a centrifugal force to it, vent the lighter-density fluid back to the annulus, and transfer the heavier-density fluid to the first pump stage. The heavier-density fluid, which is routed to the pump, has been either fully or partially degassed. Two of these devices are shown in the left-hand and center graphics of Fig. 13.3. The first device is the vortex-type separator. The produced fluid, which has already undergone some natural annular separation, is drawn into the unit through the intake ports. These can be straight intake ports, as already mentioned, or a reverse-flow-intake style. The fluid is then boosted to the vortex generator by the positive-displacement inducer. The vortex generator is generally an axial-type impeller. It imparts a high-velocity rotation to the fluid. This causes the heavier fluids (liquids) to be slung to the outer area of the flow passageway and the lighter fluids (free-gas laden) to mingle around the inner area and the shaft. The fluid then enters a stationary flow-crossover piece. The crossover has an outer annular passageway that takes the heavier-density fluids that enter it and directs them to the entrance of the pump. The lighter-density fluid that enters the inner annular passageway of the crossover is directed to the separator vents, where it exits to the casing annulus and flows up the wellbore.

Flanged Connection to Seal-Chamber Section. The bottom flange of the pump bolts to the flange of the seal-chamber-section head. It maintains axial alignment of the shafts of the two units. It also allows the floating pump shaft to engage the end of the seal-chamber-section shaft so that the axial thrust produced by the pump is transferred to the thrust bearing in the seal-chamber section.

Stages. The stages of the pump are the components that impart a pressure rise to the fluid. The stage is made up of a rotating impeller and stationary diffuser. The stages are stacked in series to incrementally increase the pressure to that calculated for the desired flow rate. A graphic of the fluid flow path is illustrated in Fig. 13.4. The fluid flows into the impeller eye area and energy, in the form of velocity, is imparted to it as it is centrifuged radially outward in the impeller passageway. Once it exits the impeller, the fluid makes a turn and enters the diffuser passageway. As it passes through this passageway, the fluid is diffused, or the velocity is converted to a pressure. It then repeats the process upon entering the next impeller and diffuser set. This process continues until the fluid passes through all stages, and the design discharge pressure is reached. This pressure rise is often referred to as the total developed head (TDH) of the pump.

A key feature for both styles of stages is the method by which they carry their produced axial thrust. Usually, the pumps that are under a 6-in. diameter are built as "floater" stages. On these, the impellers are allowed to move axially on the pump shaft between the diffusers. Contrary to the name given to this configuration, the impellers never truly float. They typically run in a downthrust position, and at high flow rates, they may move into upthrust. To carry this thrust, each impeller has synthetic pads or washers that are mounted to the lower and upper surfaces, as shown in the previous figures. These washers transfer the thrust load from the impeller through a liquid film to the smooth thrust pad of the stationary diffuser.

On 6-in. and larger pumps and on specially built smaller pumps, the impellers are usually fixed or locked to the shaft. These pumps are referred to as "fixed impeller" or "compression" pumps. In this configuration, all the thrust is transferred to the shaft and not to the diffuser. Therefore, the seal thrust bearing carries the load of all the impellers plus the shaft thrust. Particular care should be exercised in selecting the proper seal thrust bearing to match the fixed impeller pump conditions because these loads can be very high.

Performance Characteristics. The manufacturers state the performance of their pump stages on the basis one stage, 1.0 specific gravity (SG) water at 60- or 50-Hz power. A typical performance curve for a 4-in.-diameter radial-style pump, with a nominal best-efficiency performance flow of 650 B/D, is shown in Fig. 13.13. A mixed-flow style with a nominal flow rate of 6,000 B/D is shown in Fig. 13.14. In these graphs, the head, brake horsepower (BHP), and efficiency of the stage are plotted against flow rate on the x -axis. Head, flow rate, and BHP are based on test data, and efficiency is calculated on the basis of

The head/flow curve shows the head or lift, measured in feet or meters, which can be produced by one stage. Because head is independent of the fluid SG, the pump produces the same head on all fluids, except those that are viscous or have free gas entrained. If the lift is presented in terms of pressure, there will be a specific curve for each fluid, dependent upon its SG.

The dark (highlighted) area on the curve is the manufacturers recommended "operating range." It shows the range in which the pump can be reliably operated. The left edge of the area is the minimum operating point, and the right edge is the maximum operating point. The best efficiency point (BEP) is between these two points, and it is where the efficiency curve peaks. The shape of the head/flow curve and the thrust characteristic curve of that particular stage determines the minimum and maximum points. The minimum point is usually located where the head curve is still rising, prior to its flattening or dropping off and at an acceptable downthrust value for the thrust washer load-carrying capabilities. The location of the maximum point is based on maintaining the impeller at a performance balance based on consideration of the thrust value, head produced, and acceptable efficiency.

API RP11S2 covers the acceptance testing of ESP pumps. H) is a function of diameter (D) to the second power and also of rotating speed (N) to the second power. Flow (Q) is a function of diameter to the third power and also a direct function of rotating speed.

The BHP curve shows the power required to drive the stage. The power is lowest at shutoff or zero flow and increases with flow. The HP also follows the relationship that is given in Eq. 13.4 for different-sized pumps under dynamically similar conditions.

For any particular-diameter-pump series, there is generally an overlap region between the radial and mixed-flow styles. A typical relationship of a family of similar-diameter stages is shown in Fig. 13.15. Notice that each style increases in efficiency as the flow rate increases, until the efficiency peaks and begins dropping off.

The component located below the lowest pump section and directly above the motor, in a standard ESP configuration, is the seal-chamber section (Fig. 13.16). API RP11S7 gives a detailed description of the design and functioning of typical seal-chamber sections. RP11S7. The seal-chamber section is basically a set of protection chambers connected in series or, in some special cases, in parallel. This component has several functions that are critical to the operation and run-life of the ESP system, and the motor in particular.

Axial Thrust Bearing. This bearing carries all of the axial thrust produced by the pump and seal-chamber section. Generally, sliding-shoe hydrodynamic types are used for this application because of their robustness and ability to function totally immersed in lubricating fluid. It is composed of two main components: a stationary pad and a rotating flat disk. The stationary part has pads finished to a very close flatness tolerance, connected to a base by a thin pedestal or flexible joint. The rotating disk is also finished to a very close flatness tolerance. Several different bearing designs are shown in Fig. 13.22. They represent standard-style cast bearings for normal applications and machined bearings for intermediate- and high-load applications.

The shaft has to transmit, from the motor to the pump, the entire torque required by the equipment for its application. This not only includes the stabilized running torque but also the short-term torque spikes caused by unit startup and intermittent pump loads. Because the diameter of the shaft is constrained because of the maximum diameter of the unit, materials of differing mechanical properties must be used to provide different load capabilities. These materials must also provide protection from corrosive wellbore fluids.

The thrust-bearing performance is a function of the load that is transferred to it and the viscosity of its lubricating oil. The load transmitted from the pump can be calculated on the basis of the pump geometry and the TDH produced for the application. For "floater" pumps, the shaft load is always down and is equal to the cross-sectional area of the top of the shaft multiplied by the discharge pressure of the pump (Pdischarge) minus the cross-sectional area of the bottom of the shaft multiplied by the pump intake pressure (PIP). For "fixed" impeller pumps, the load is equal to the shaft force, as just calculated, plus the summation of all the impeller thrust forces. The impeller thrust forces can be roughly calculated, as previously described in the pump-stage section, or obtained from the pump manufacturer.

Revolutions per Minute (RPM). The rotational speed or RPM of the motor at its application load point is very important in determining the operating point or output of the pump. The pump-performance curve used in determining the head and flow output of the pump for its application is based on a pump-motor speed of 3,500 RPM. If the RPM varies from 3,500, the pump flow will vary with the ratio of the speed, and the flow rate will vary with the ratio of the speed squared. (See Eqs. 13.1 and 13.2.) Once again, by knowing the percent of nameplate amps, the motor speed can be read from the motor characteristic curve. Even though this RPM change is usually small, it can still impact the final motor and pump operating point for a particular application. When the pump-performance point is modified, because of the motor RPM, the pump head and flow rate change; therefore, the load on the motor is changed. Determining the final pump operating point and motor loading point becomes an iterative process.

Motor Lead Extension (MLE). The motor lead extension cable, also referred to as the motor flat, is a specially constructed, low-profile, flat cable. It is spliced to the lower end of the round or flat main power cable, banded to the side of the ESP pump and seal-chamber section, and has the male termination for plugging or splicing into the motor electrical connection. Because of its need for low profile, it requires compact construction. It generally has a thin layer of high-dielectric-strength polyamide material wrapped or bonded directly to the copper conductors. This allows for a thinner layer of insulation material, allowing for a lower profile. The MLE is generally selected on the basis of equipment: casing clearance and the voltage capacity requirement.

Control Module. These are solid-state devices that offer basic functions necessary to monitor and operate the ESP in a reliable manner. The unit examines the inputs from the CPT and other input signals and compares them with preprogrammed parameters entered by the operator. Some of the functions include overload and time-delayed underload protection, restart time delay, and protection for voltage or current imbalance. Additional external devices can be connected, which provide for downhole pump intake pressure protection, downhole motor temperature protection, surface tank high/low level controls, line pressure switches, and others.

VSCs used with ESPs should be designed for the specific requirements of the downhole ESP motor and pump. This is because of the unique design and characteristics of the downhole centrifugal pump and submersible motor as compared to their surface counterparts. Generally, the VSC is designed to provide a constant volts/hertz output through a broad range of frequency variations. The magnetic flux that is generated in the stator of the submersible motor and passes through the rotors is directly proportional to the voltage and inversely proportional to the frequency of the applied power. The result is a constant magnetic flux density in the motor. Because the output torque of the motor is proportional to the magnetic flux density, the motor is a constant-torque variable-speed device. Also, because of its low inertia characteristics and unique rotor design, it does not have the same high-operating-speed restrictions as a typical surface induction motor. Therefore, a VSC is typically applied to frequencies from 30 to 90 Hz, with its minimum and maximum frequencies restricted only by the mechanical limitations of the downhole ESP equipment.

Because of the relationship of the performance of a centrifugal pump to its rotational speed (Eqs. 13.2 through 13.4), the VSC allows for wider flexibility of the downhole ESP system. The effect on pump operation is shown in Fig. 13.34. This is the same pump that is represented in the 60-Hz fixed-speed performance curve of Fig. 13.14. This allows the designer to select the flow rate and speed of the system on the initial design. For this pump stage, it can be operated between 1,800 B/D at 30 Hz (minimum recommended operating point) and 10,200 B/D at 90 Hz (maximum recommended operating point). The benefits of VSC usage are discussed next.

Broadened Application Range. On fixed-speed operation, a pump stage has a recommended minimum and maximum flow rate. Beyond these points, the pump can operate in a detrimental run-life or reliability area. By operating at reduced frequency, the minimum recommended operating point is reduced, and, at higher frequencies, the maximum operating point is increased. This allows the application of ESPs in low-productivity-index (PI) wells and higher flow rates to be obtained from small bore casings. It also allows a limited inventory of pumps to be applied over a broader flow range.

Maximize Well Production. If the well PI is greater than that for the original design, either through data error or changing wellbore parameters, the ESP operating point can be increased with a VSC. The HP rating of the motor limits the frequency increase. Remember, the HP load from the pump increases with the cube of the frequency ratio, and the HP capability of the motor increases directly to the speed ratio. Therefore, the designer must consider using an oversized motor if there is a potential need of higher flow rates.

Minimum Well Production. If the well PI is lower than that for the original design, the ESP operating point can be decreased with the VSC. The TDH of the pump is the limiting factor on the minimum VSC frequency. The produced head of the pump decreases with the square of the frequency ratio. Therefore, the designer must consider initially oversizing the pump lift, if there is a potential for reduced-frequency operation.

Pump Intake or Casing Annulus Pressure. This information provides wellbore static pressure and the well flowing pressure at the production rate. If the measurement is sensitive enough, it can also provide excellent well drawdown information.

Pump Discharge Pressure. This parameter provides a reading on the discharge pressure of the pump. This reading and the pump intake pressure provide a measurement of the TDH of the pump. Comparing this value to the design TDH, hydraulic performance of the pump can be monitored and continually evaluated. Additionally, for gassy and/or viscous fluids, pump-performance correction factors can be established or verified for that particular wellbore condition.

Pump Discharge Temperature. This measurement provides the temperature of the discharge fluid from the pump. The production fluid is heated as a result of the heat rejected by the motor and pump inefficiencies. The fluid heat rise through the pump can be used to calculate the fluid volumetric increase and the viscosity change of the fluid. Once again, sudden spikes or longer-term changes can provide warnings of potential problems.

Downhole Flow Rate. Downhole flowmeters are available that provide flow-rate measurements from the pump discharge. This is an excellent tool, when compared to the surface flow rate, for evaluating ESP performance and warning of potential problems. Because surface flow rate is not generally continuously monitored, this can be a piece of information for enhanced ESP protection. In multiphase-fluid (gassy) applications, the selection and calibration of the flowmeter is important because of the difficulty in accurately measuring this fluid.

Packers are used with ESP systems when there is a need to isolate the annular area above the ESP and/or provide a positive barrier between the pressurized wellbore fluid and the area above the packer. Isolating the area above the packer is done to segregate two separate zones or prevent or reduce the rate of wellbore fluid corrosion damage to the casing. With a deep-set packer, operational precautions must be observed to prevent damage to the ESP system. With a deep-set packer, the volume contained between the packer and pump intake is usually small. Upon startup, the ESP can evacuate this volume quickly, causing a sudden drop from wellbore static to flowing pressure. This causes sudden decompression to the cable and internal volumes of the seal-chamber section and motor, especially if they have been saturated with solution gas. This decompression can cause expansion and insulation damage to the cable. If it is severe enough, it can result in extensive expulsion of mot