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Petroleum Engineering Handbook

Larry W. Lake, Editor-in-Chief

Volume III – Facilities and Construction Engineering

Kenneth E. Arnold, Editor

Chapter 6 – Pumps

Maurice I. Stewart Jr., Stewart Training Co.

Pgs. 231-259

ISBN 978-1-55563-116-1
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Fluid Principles and Hydraulics

Types of Fluids

Pumps are used to move fluids, which include liquids, dissolved gases, and solids. Gases include dissolved air and hydrocarbon vapors. Solids include sand, clay, corrosion byproducts, and scale. The most common types of liquids pumped in upstream operations are crude oil, condensate, lube oils, glycols, amines, and water. Each fluid has different physical properties that must be taken into consideration when sizing and selecting a pump. The most important physical properties are suction pressure, specific gravity, viscosity, vapor pressure, solids content, and lubricity.

Types of Pumps

A pump can be defined as "a mechanical device that adds energy to a fluid to increase its flow rate and static pressure."[1] This process can be accomplished with positive-displacement pumps or kinetic-energy pumps.

Positive-Displacement Pumps. Positive-displacement pumps add energy to a fluid by applying force to the fluid with a mechanical device such as a piston, plunger, or diaphragm. There are two types of positive-displacement pumps: reciprocating and rotary. Reciprocating pumps use pistons, plungers, or diaphragms to displace the fluid, while rotary pumps operate through the mating action of gears, lobes, or screw-type shafts.

Kinetic-Energy Pumps. Kinetic energy (energy associated with motion) is added to a liquid to increase its velocity and, indirectly, its pressure. Kinetic-energy pumps operate by drawing liquid into the center of a rapidly rotating impeller. Radial vanes on the impeller throw the liquid outward toward the impeller rim. As liquid leaves the impeller, it comes in contact with the pump casing or volute. The casing is shaped to direct liquid toward a discharge port. The casing slows the liquid and converts some of its velocity into pressure. There are three classes of kinetic-energy pumps: centrifugal, regenerative-turbine, and special-effects pumps. The centrifugal-pump category includes radial-, axial-, and mixed-flow designs. Centrifugal pumps account for more than 80% of pumps used in production operations because they exhibit uniform flow, are free of low-frequency pulsations, and are not subject to mechanical problems. Fig. 6.1 illustrates pumps commonly used in upstream production operations.

Pumping-System Design

Designing any pumping service involves three major activities: process design, mechanical design, and vendor selection.

Process Design. The first step in process design is to obtain a design flow rate. The design flow rate should be selected after considering all flow variations, such as startup conditions, future expansion, and maximum anticipated flow. The next step is to determine the liquid properties critical to pump design. These properties include specific gravity, temperature, viscosity, pour point, etc. Values are required at pumping conditions and, in some cases, at ambient conditions as well. The next step is to calculate available suction conditions such as rated suction pressure, maximum suction pressure, and net positive suction head available (NPSHA). (See Sec. 6.2.3 for information on NPSHA.) Once the available suction conditions have been established, the effect of the selected control system on pump performance requirements must be determined (see Sec. 6.3.10). The next step is to calculate the minimum discharge-pressure requirements of the pump. The last step is to calculate the total dynamic head (TDH) at the specific gravity corresponding to rated pumping temperature.

Mechanical Design. The first step in mechanical design is to determine the design pressure and temperature required for the pump and its associated piping. Once this is done, a pump type and materials of construction are selected. The next step is to determine the sparing (backup) requirements, the need for parallel operation and control-system details. Then select a shaft-seal type and determine the requirements for an external flushing or sealing system and estimate the power requirements and choose a driver (motor, engine, or turbine) for the pump. Lastly, document the design by including calculations, studies, design specifications, utility requirements, and estimate summary.

Vendor Selection. Factors that have the greatest influence on the selection of the most cost-effective pump type include capacity, TDH, maintenance, viscosity, and capacity control. Within the general type selections, a particular construction style is most influenced by discharge pressure, NPSHA, fluid temperature, and space and weight limitations.

Hydraulic Principles


Hydraulics deals with the mechanical properties of water and other liquids and the application of these properties to engineering. Hydraulics is divided into two areas: hydrostatics (fluids at rest) and hydrodynamics (fluids in motion).


A liquid has a definite volume when compared to a gas, which tends to expand to fit its container. When unconfined, a liquid seeks the lowest possible level. Because of its fluidity, a liquid will conform to the shape of its container.

Pressure. The pressure existing at any point in a liquid body at rest is caused by the atmospheric pressure exerted on the surface plus the weight of the liquid above that point. This pressure is equal in all directions.

Temperature. For most liquids, an increase in temperature decreases viscosity, decreases specific gravity, and increases volume. Temperature affects the type of pump construction, material selection, corrosive properties of the fluid, and the pump’s flange pressure/temperature rating.

Air Properties. Air is a mixture of oxygen, nitrogen, and other compounds. The standard pressure of air is defined at 60°F, 36% relative humidity, and sea level. The weight of a column of air above the Earth’s surface at 45° latitude and sea level is 14.696 psia (29.92 in. of mercury). Atmospheric pressure decreases by approximately 0.5 psi for each 1,000 ft of elevation above sea level.

Head. The relationship of head to pressure is expressed as

where h = height of the fluid column above a reference point, and p = pressure. The head of liquid is not related to the area occupied by the liquid. Fig. 6.2 illustrates types of head.

where γ = specific gravity of the liquid, ρf = density of the liquid being pumped, and ρw = density of water at standard conditions of temperature and pressure.

It is important to realize that although the heads of different liquids are the same, their pressures are different because of the differences in specific gravities. For example, assume three 100-ft-tall tanks filled with gasoline, water, and molasses, respectively. The pressure measured at the bottom of each tank is different because of the differences of specific gravities of gasoline (0.75), water (1.0), and molasses (1.45).

Centrifugal-Pump Considerations. Operating pressure is expressed in feet of the liquid that is being pumped. Suction and discharge pressures are expressed as suction head and discharge head, respectively. Pressures are expressed in feet of head, because it is more important to know how much a pump can raise the liquid it is pumping, rather than the amount of pressure the pump is adding to the liquid.

Positive-Displacement Pump Considerations. Operating pressures are almost always expressed in terms of pressure (psi).

Static Head, Static Lift, and Submergence Terminology. Fig. 6.3 illustrates the relationship between static head, static lift, and submergence. Static head is the vertical distance between a liquid level and a datum line, when the supply is above the datum. Static lift is the vertical distance between a liquid level and a datum level, when the datum is above the liquid. Datum line is the centerline of the pump inlet connection, or the horizontal centerline of the first-stage impeller in vertical pumps.

Theoretical Lift. A pump that develops a perfect vacuum at its suction end can lift a column of water 34 ft. This vertical distance is called theoretical lift. The pressure to lift the liquid comes from atmosphere pressure. At sea level, atmosphere pressure is approximately 14.7 psia.

Actual Suction Lift. Because a perfect vacuum is never achieved and because some lift is lost to friction in the suction line, the maximum actual suction lift for a positive-displacement pump is approximately 22 ft. The maximum actual suction lift for a centrifugal pump is approximately 15 ft when pumping water from an open air tank. Positive-displacement pumps can operate with lower suction pressures or high suction lifts because they can create stronger vacuums. Suction lift will be greater if the pressure in a closed tank is greater than atmospheric pressure.

Submergence. Submergence is often confused with either suction static head or static lift. For vertical pumps, submergence relates the liquid level to the setting of the pump. For horizontal pumps, submergence relates to the height of liquid level necessary in the source vessel or tank to prevent the formation of vortexing and the resulting flashing of vapors in the pump suction.

Vapor Pressure. As the pressure on a liquid is decreased, there is a tendency for the bubbles of vapor to be liberated. The vapor pressure of a liquid is the pressure at which the first bubble of vapor appears at a given temperature. At 60°F, the vapor pressure of water is 0.3 psia (0.7 ft). At 212°F, the vapor pressure of water is 14.7 psia (34 ft). Fig. 6.4 illustrates the vapor pressure of water for various temperatures. For other fluids, refer to standard references (e.g., Hydraulic Institute Engineering Data Book[1]).

Suspended Solids. The amount and type of suspended solids entrained in the liquid can affect the characteristics and behavior of that liquid. Increased concentrations of solids increase the specific gravity, viscosity, and abrasiveness of a liquid. The type and concentration of suspended solids can affect the style of pump selected and the materials of construction. Suspended solids also affect the selection of impeller design in centrifugal pumps, which in turn affects the wear rate, efficiency, and power consumption.

Dissolved Gases. Small amounts of dissolved gases have little effect on flow rate or other pumping requirements. If large amounts of gas enter the liquid through piping leaks or as a result of vortexing in vessels, the specific gravity of the liquid will decrease. Dissolved gases can also reduce the amount of NPSHA at the pump suction. (See Sec. 6.2.3 for a discussion of NPSHA).

Viscosity. Viscosity offers resistance to flow because of friction within the fluid. Viscosity levels have a significant impact on pump type selection, efficiency, head capacity, and warm-up. High-viscosity liquids decrease a centrifugal pump’s efficiency and head performance, while increasing the power requirements. The viscosity of all liquids varies with temperature. For viscosities of liquids, refer to standard industry references (e.g., Hydraulic Institute Engineering Data Book[1]).

Corrosivity. The corrosive nature of the fluid being pumped has a bearing on pump type selection, materials of construction, and corrosion allowance. Special mechanical seals and flushing arrangements may be required.


Hydrodynamics is the study of fluids in motion. Bernoulli’s equation states that


where v = average velocity of the liquid in the pipe, g = acceleration of gravity, p = pressure, ρ = density, Z = height above a datum, and hf = friction loss between points 1 and 2. Subscripts 1 and 2 refer to locations along a pipe. An examination of each of the terms in Eq. 6.3 provides a better understanding of the general equation for modeling a pumping system.

Velocity Head. Velocity head is the potential energy that has been converted to kinetic energy. Velocity head can be expressed as


or RTENOTITLE....................(6.5)

where Q = flow rate, and d = inside pipe diameter.

The velocity head increases the amount of work required of a pump. The velocity head is usually not included in actual system calculations when piping velocities are kept within the prescribed limits of 3 to 15 ft/sec. The velocity head is included in the total dynamic head on the centrifugal-pump curves.

Pressure Head. The energy contained in the liquid is expressed as pressure head and expressed as p/ρ in Eq. 6.3.

Elevation Head. The energy contained in the liquid as a result of its elevation relative to a datum is called the elevation head and is expressed as Z in Eq. 6.3.

Head Losses. Head losses are potential energy that has been lost because of frictional resistance of the piping system (pipe, valves, fittings, and entrance and exit losses). Unlike velocity head, friction head cannot be ignored in system calculations. Head loss values vary as the square of the flow rate. Head losses can be a significant portion of the total head. See the chapter on Piping and Pipelines in this volume of the Handbook for a discussion of pressure drop (i.e., head losses in piping).

Control Losses. Control losses occur on the discharge side of a centrifugal pump that has been equipped with a backpressure valve to control flow rate. As the liquid flows through the control valve, energy is lost. Next to static head, control losses are frequently the most important factor in calculating the pump’s total dynamic head. For pump applications, control losses are treated separately from head losses, even though they are included in the hf term in Eq. 6.3.

Acceleration Head. Acceleration head is used to describe the losses associated with the pulsating flow of reciprocating pumps. Theoretically, acceleration head should be included in the hf term of Eq. 6.3. Hydraulic Institute Engineering Data Book[1] discusses the calculation of acceleration head.

Total Dynamic Head. TDH is the difference between the pumping system’s discharge head and suction head. It is also equal to the difference in pressure-gauge readings (converted to feet) across an existing operating pump (discounting velocity head).

Suction Head. Suction head is defined as the sum of the suction-vessel operating gauge pressure (converted to feet), the vertical distance between the suction-vessel liquid level and the pump reference point, less head losses in the suction piping [discounting change in velocity,


and acceleration head]. Suction head can be expressed as


which can be reduced to


where Hs = suction head of liquid being pumped, p1 = suction-vessel operating pressure, H1 = height of liquid suction vessel above pump reference point, and pf1 = pressure drop resulting from friction in the suction piping.

Discharge Head. Discharge head is defined as the sum of the discharge-vessel operating gauge pressure (converted to feet), the liquid level in the discharge vessel above the pump reference point, pressure drop because of friction in the discharge piping, and control losses (discounting velocity head). It can be expressed as


which can be reduced to


where Hd = discharge head of liquid being pumped, p2 = discharge-vessel operating pressure, H2 = operating or normal height of liquid in the discharge vessel above the pump reference, pf2 = pressure drop resulting from friction in the discharge piping, and Pc = discharge flow-control-valve losses.

Calculating TDH. The pump TDH is the difference between the suction and discharge heads.


which can be substituted as


where Htd = total dynamic head required of a pump.

Net Positive Suction Head (NPSH). NPSH is defined as the total suction head in feet of liquid (absolute at the pump centerline or impeller eye) less the vapor pressure (in feet) of the liquid being pumped.

Net Positive Suction Head Required. Net positive suction head required (NPSHR) is defined as the amount of NPSH required to move and accelerate the liquid from the pump suction into the pump itself. It is determined either by test or calculation by the pump manufacturer for the specific pump under consideration. NPSHR is a function of liquid geometry and the smoothness of the surface areas. For centrifugal pumps, other factors that control NPSHR are the type of impeller, design of impeller eye, and rotational speeds. NPSHR is determined on the basis of handling cold water. Field experience coupled with laboratory testing have confirmed that centrifugal pumps handling gas-free hydrocarbon liquids and water at elevated temperatures will operate satisfactorily, with harmless cavitation and less NPSHR than would be required for cold water.

Net Positive Suction Head Available. NPSHA must be equal to or greater than NPSHR. If this is not the case, cavitation or flashing may occur in the pump suction. Cavitation occurs when small vapor bubbles appear in the liquid because of a drop in pressure and then collapse rapidly with explosive force when the pressure is increased in the pump. Cavitation results in decreased efficiency, capacity, and head and can cause serious erosion of pump parts. Flashing causes the pump suction cavity to be filled with vapors and, as a result, the pump becomes vapor locked. This usually results in the pump freezing up, which is called pump seizure.

NPSHA is not a function of the pump itself but of the piping system for the pump. It can be calculated from


where pA = atmospheric pressure and pva = liquid vapor pressure at pumping temperature.

NPSHA decreases with increases in liquid temperature and pipe friction losses. Because pipe friction losses vary as the square of the flow, NPSHA also varies as the square of the flow. Thus, NPSHA will be the lowest at the maximum flow requirement. Accordingly, it is important to recognize the need for calculating NPSHA (and NPSHR) at maximum flow conditions as well as maximum fluid temperature, not just at design conditions. Unless subcooled, a pure-component hydrocarbon liquid is typically in equilibrium with the vapors in a pressure vessel. Thus, increases in the vessel operating pressures are almost fully offset by a corresponding increase in the vapor pressure. When this occurs,


NPSH Margin. The NPSH margin is NPSHA less the NPSHR. The Hydraulic Inst. recommends an NPSH margin of 3 to 5 ft. [1]

When a new system offers insufficient NPSH margin for optimum pump selection, either the NPSHA must be increased, the NPSHR must be decreased, or both. To increase the NPSHA, one can raise the liquid level, lower the elevation of the selected pump, change to a low-NPSHR pump, or cool the liquid. To reduce the NPSHR, one can use different design impellers or inducers or use several smaller pumps with lower NPSHRs in parallel.

When an existing pumping system exhibits insufficient NPSH margin, it is too late to use these solutions without going through an expensive change. Most of these problems can be traced to suction flow restrictions (orifice plates, plugged strainers, partially closed valves, etc.) and inadequate source-tank liquid levels.

Power Requirements. Once the TDH has been calculated, the power requirements can be determined with


For kinetic-energy pumps,


For positive-displacement pumps,


where PB = brake horsepower, and e = the pump efficiency factor obtained from the pump manufacturer. For electric-motor-driven pumps, the energy consumption can be estimated with


Centrifugal Pumps


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.

Basic Classifications

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.

Impeller Types

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.

Number of Impellers

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.

Impeller Axial Loading

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.

Impeller Radial Loading

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

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
where Ns = pump specific speed, N = pump rotative speed, q = pump capacity, and Htd′ = TDH per stage at the BEP.

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.

Pump-Performance Curves

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.

System Head Curves

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.

The constant portion represents the static head difference between the suction and the discharge at zero flow and is equal to


The variable portion represents the head required to overcome friction as a result of flow. It varies as the square of the flow and is equal to


where pf1 = pressure drop resulting from friction in the suction piping, pf2 = pressure drop resulting from friction in the discharge piping, and Pc = discharge flow-control-valve losses.

Regulation of Flow Rate

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.

Backpressure Valves

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.

Minimum-Flow Recirculation Valve

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.

Changing Performance

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.

Variable-Speed Control

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%.

Affinity Laws

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.

For a change in pump speed, the following changes in pump performance can be determined:



and RTENOTITLE....................(6.24)

where N1 = old speed, and N2 = new speed. For a change in diameter, the following performance changes can be determined:



and RTENOTITLE....................(6.27)

where D1 = old diameter, and D2 = new diameter. For a change in both diameter and speed, the following changes in pump performance can be determined:



and RTENOTITLE....................(6.30)

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%.

The efficiency is assumed to be constant in all the previous calculations. Fig. 6.10 illustrates a graphic example of reduced operating parameters because of speed reducers.

Pump Priming

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.

Installation Considerations

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.

Suction Piping. Suction piping is more important than discharge piping.

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.

Basket Strainers. Conditions may dictate that permanent strainers be installed in the suction piping. If permanent strainers are not required, temporary cone-type strainers should be installed at least for initial startups. Basket strainers should have at least 150% flow-area screens.

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.

Positive-Displacement Pumps


Positive-displacement pumps were developed long before centrifugal pumps. Liquid is positively displaced from a fixed-volume container. Positive-displacement pumps are capable of developing high pressures while operating at low suction pressures. They are commonly referred to as constant-volume pumps. Unlike centrifugal pumps, their capacity is not affected by the pressure against which they operate. Flow is usually regulated by varying the speed of the pump or by recycle. Positive-displacement pumps are divided into two groups: rotary and reciprocating pumps.

Rotary Pumps

Rotary pumps are normally limited to services in which the fluid viscosity is very high or the flow rate too small to be handled economically by other pumps. Rotary pumps are commonly used to circulate lube oils through engines, turbines, reduction gears, and process-machinery bearings. Rotary pumps displace a fixed quantity of fluid for every revolution of the driver shaft. They have different pumping elements such as vanes, lobes, gears, and screws. Fig. 6.11 illustrates three (internal gear, external gear, and screw) of the most commonly used rotary pumps in production operations.

Most manufacturers rate rotary pumps by capacity (i.e., throughout). Capacity is the total liquid displacement of the pump less slip. Slip is the quantity of fluid that leaks from the higher-pressure discharge to the lower-pressure suction. Slip occurs because all rotary pumps require clearances between the rotating elements and pump housing. These clearances provide a leak path between the discharge and suction sides. A pump with large clearances, because of machining tolerances or wear, exhibits a proportionally larger slip. Rotary pumps cannot move nonlubricating fluids such as water or fluids containing hard or abrasive particles. Rotary pumps can move large quantities of air or vapor for short periods of time without losing prime.

Rotary pumps are self-priming but are not designed to run dry for long periods. For best operation, there must be enough fluid at the suction port to keep the pumping chamber completely filled.

Fig. 6.12 illustrates the relationship between speed, volumetric efficiency, and displacement of a rotary positive-displacement pump. The principles of operation of some of the more common types of rotary pumps are described next.

Sliding Vane. A set of vanes is mounted in a rotor in which the vanes slide in and out of the rotor. The rotor is mounted off center in the casing. As the vanes rotate past the suction port, they slide out of the rotor while maintaining constant contact with the casing. Springs or sealer rings help hold the vanes against the casing, thus the vanes make a close seal, or fit, against the casing wall. Trapped fluid is forced from the suction port to the discharge port.

The sliding-vane design is capable of delivering medium capacity and head. They deliver a constant flow rate for a set rotor speed. They work well with low-viscosity fluids and are somewhat self-compensating for wear. They are not suitable for use with highly viscous fluids (thicker fluids interfere with the sliding action of the vanes). A large wear area results from the friction fit between the vanes and the cylinder.

Flexible Vane. The flexible vane is similar to the sliding vane except that the vanes are generally a soft, pliable material and are integral with the rotor. As the rotor turns, the vanes bend and conform to the eccentric shape of the cylinder. They are simple, inexpensive, and are capable of developing a vacuum. They should not be allowed to run dry and should be used only with low-temperature fluids and in low-head applications.

External Gear. The external gear consists of two equal-sized meshing gears, one a driver and the other an idler, that rotate inside a housing. As the gears unmesh at the suction side of the pump, a vacuum is formed. Pressure forces the fluid into the pump where the fluid is carried between the gear teeth and the case to the discharge port. At the discharge, the meshing of the gear teeth creates a boundary that prevents the fluid from returning to the suction. Gear pumps operate equally well when driven in either direction. Precautions should be taken to ensure that the shaft rotation is correct when special features, such as built-in relief valves or a bleed back of the shaft seal, are used.

There are also models that use multiple sets of gears on one shaft to produce more capacity. External-gear pumps are compact in size and can produce high pressures. They are well suited for highly viscous fluids. They are easily manufactured in a broad range of materials to ensure compatibility with the pumped fluids. Because of their close tolerances, they are limited to clean-fluid applications.

Internal Gear. The internal-gear pump is similar in principle to the external gear except the drive shaft turns a ring gear with internal teeth. The external gear tooth (idler) rotates on an offset center and meshes with the drive gear through only a segmental arc of rotation. A fixed crescent-shaped filter occupies the space between internal- and external-gear-tooth tips opposite the mesh point. As the gear teeth disengage at the input port, fluid enters and is trapped in the tooth space of each gear and is carried to the discharge port. The meshing of the two gears and the elimination of the tooth space forces fluid from the pump.

Lateral gear pumps are used in low-head applications. They are limited to a maximum backpressure of 100 psi and require a pressure-relief valve on the discharge side. Because small clearances exist, they cannot handle liquids that contain solids. The manufacturer should always be consulted before any gear pump is used with fluid-handling solids.

Lobe. Lobe pumps operate in the same manner as gear pumps except the rotating elements have two, three, or four lobes instead of gear teeth. Lobes cannot drive each other, so timing gears are used. The lobes never come into contact with each other so the pump can be allowed to run dry. Lobes are used where product integrity must be maintained and in applications in which liquids are shear sensitive. The large volume created between the casing and lobes allows many products to be pumped without damaging the product itself. A major advantage is that there is no metal-to-metal contact between the lobes, thus the possibility of traces of iron, steel, or other pump-construction materials ending up in the product because of wear is greatly reduced. On the other hand, they are more expensive than gear or vane pumps and are difficult to repair and maintain.

Screw. Screw pumps can be single-rotor (progressive cavity) or multiple-rotor (intermeshing) design. Screw pumps are relatively high-speed pumps but, because of the reversal of flow required to enter the suction passage, NPSH can often be a problem. Screw pumps are used for high-head applications; thus, they are the most common rotary-pump type in use in producing operations.

Single Screw. In the single-screw design, the fluid is trapped between the treads of a rotating screw and the treads of the internal stationary element. These pumps are used for viscous liquids and liquids with high solids content. They can produce significant suction lift and relatively high pressures. They can handle fluids ranging from clean water to sludges without changing clearances or components. On the other hand, they are expensive, bulky, and difficult to maintain, and replacement parts are expensive.

Multiple Screw. In the multiple-screw design, the fluid flows between a central drive screw and one or more idler screws in a close-fitting housing. In two-screw pumps, both shafts are driven with timing gears. In three-screw pumps, the screw treads are cut so one screw can drive the other two. The rotation of the screws produces a vacuum at the inlet, moves the fluid through the pump, and delivers the fluid to the discharge. In small sizes, they are used to supply lubricating oil to engines and industrial machinery. In intermediate sizes, they are used in office buildings as a source of hydraulic energy to operate elevators. In large sizes, they are used to load and unload barges and tankers.

Reciprocating Pumps

Reciprocating pumps move liquid by means of a constant back-and-forth motion of a piston, plunger, or diaphragm within a fixed volume or cylinder. Reciprocating pumps can handle viscous and abrasive fluids. They are low-speed machines when compared with centrifugal and rotary pumps. They offer higher efficiencies, generally 85 to 94%, thus they require less horsepower. Reciprocating pumps are best suited for high-pressure and low-volume applications. They frequently require pulsation dampeners because of the pulsating nature of the flow. They have higher installed costs (usually offset by higher efficiencies) and higher maintenance costs than centrifugal or rotary pumps.

Plunger and Piston Pumps. In plunger pumps, a plunger moves through a stationary packed seal and is pushed into and withdrawn from a liquid cavity. In piston pumps, a piston moving back and forth within a liquid cavity pushes the fluid from the cylinder. Movement of either the plunger or piston creates an alternating increase and decrease of flow. As the plunger or piston moves backward, the available volume in the cylinder increases and a suction valve opens to allow the liquid to enter the cylinder through a one-way suction valve. As the plunger or piston moves forward, the volume available in the cylinder decreases, the pressure of the liquid increases, and the liquid is forced out through a one-way discharge valve.

Efficiencies remain high regardless of head or speed (tend to decrease slightly with increasing speed). Because reciprocating pumps run at lower speeds than centrifugal or rotary pumps, they are better suited for handling viscous liquids. They are capable of producing high pressures and large capacities and are self-priming. On the other hand, they require more maintenance because of the large number of moving parts. They are heavier in weight and require more floor space than centrifugal or rotary pumps. In addition, they are poor at handling liquids containing solids that tend to erode valves and seats. Plunger and piston pumps require larger NPSHs because of pulsating flow and pressure drop through the valves. As a result of pulsating flow, they require special attention to suction- and discharge-piping design to avoid both acoustical and mechanical vibrations.

Diaphragm Pumps. Fig. 6.13 shows a typical fluid (gas-, air-, or liquid-powered) diaphragm pump. Its principle of operation is similar to plunger and piston pumps except that, instead of a plunger or piston, there is a flexible pulsating diaphragm that displaces the liquid. Varying power-fluid pressure on one side of the diaphragm causes the diaphragm to deflect alternatively drawing liquid into the pump-side chamber or discharging the liquid from the pump-side chamber. Diaphragm pumps are capable of pumping liquids that are viscous, erosive, corrosive, or that contain large amounts of solids. In addition, diaphragm pumps are self-priming, can run periodically without liquids, and are inexpensive to repair because they have no stuffing box and have few moving parts.

Diaphragm pumps are limited to small flow rates (90 gal/min), moderate discharge pressures, and moderate temperatures. They require frequent maintenance and exhibit fatigue failure with time. Leaks can cause a hazard by mixing power fluid with the process fluid. Gas-/air-powered diaphragm pumps are commonly used as sump pumps.

It is possible to use a diaphragm to power a plunger or piston pump. This type of pump is often used for chemical injection because it is well suited for low volume and large-head applications, and the speed can be controlled by a throttling valve on the power fluid.

Reciprocating-Pump Performance Considerations

Reciprocating pumps are constant-volume pumps. Variations in discharge pressures do not affect flow rate. Because these pumps continue to deliver the same capacity, any attempt to throttle the discharge flow may overpressure the pump casing and/or the discharge piping. Thus, no reciprocating pump should ever be started or operated with the discharge block valve closed. Flow is regulated by speed.

Capacity. The capacity of a reciprocating pump is the cylinder displacement less slip. For a single-acting cylinder, cylinder displacement can be determined from


For double-acting cylinders, the cylinder displacement can be determined by


where s = cylinder displacement, A = plunger or piston area, a = piston-rod cross-sectional area, LS = stroke length, N = speed, and m = number of pistons or plungers.

Slip is the loss of capacity as a percentage of the cylinder displacement because of volumetric efficiency, stuffing-box losses, and valve losses. Volumetric efficiency (not to be confused with mechanical efficiency) is normally 95 to 97%. Efficiency is also reduced when pumping a light hydrocarbon that has some degree of compressibility.

The pump capacity can be determined from


where q = pump capacity.

Speed. Speed is the primary factor that determines both the capacity of a reciprocating pump and its maintenance costs. Running at high speeds shortens packing life and increases acceleration and deceleration forces on all moving components. Operating below the maximum "rated" speed may be advantageous when the pump is operated unattended, when there are no spares and no standby, when there is a high penalty for down time, when unit maintenance is poor, when long life is desired, and when the NPSH margin is low. Operating at the maximum rated speeds requires clean, cool fluids; excellent piping layout with rigidly fixed piping; good NPSH margin; solid foundation; well-designed suction and discharge pulsation dampeners; and good maintenance. Whenever it becomes necessary to operate above the maximum rated speeds, very close attention should be given to all design, operation, and maintenance details.

Installation Guidelines

If positive-displacement pumps are properly installed and operated, satisfactory performance can be realized for a long time. These pumps are manufactured in a variety of designs for many different services. Each manufacturer’s instructions should be followed carefully for specific machines or application equipment. The following discussion relates to general installation guidelines for positive-displacement reciprocating pumps.

Foundations and Alignment. Most pump foundations are constructed of reinforced concrete. The pump and driver are bolted to a cast iron or steel base plate, which is secured to the concrete foundation with anchor bolts. Small pumps need a foundation large enough to accommodate the base-plate assembly. Large pumps require a foundation that is three to four times the weight of the pump and driver.

Anchor-Bolt-Sleeve Installation. Each anchor bolt is fitted with a washer and passed through a pipe sleeve that has a diameter three to four times greater than the bolt. The bolt-sleeve unit is set into the concrete at the predetermined base-plate hole positions. The flexibility in the sleeve washer unit allows minor adjustments to be made in the bolt position before final tightening even after the concrete foundation has set.

Metal-Shim Adjustments. Metal shims are used to position the pump on the foundation. Adjustments are made until the pump shaft and flanges are completely level. Alignment between the pump and driver is then adjusted before connecting the pump to the suction and discharge lines. The latter should have been aligned during the initial positioning of the base plate.

Grouting. Because of pipe strain, the entire pump assembly should be rechecked for alignment once the piping has been securely bolted. If the drive alignment has not been changed by bolting the piping, the space between the base plate and concrete foundations is filled with grouting. Grouting should be sufficiently fluid to fill all the available space under the base plate.

Operating-Temperature Considerations. It is essential that the alignment between the piping, pump, and driver not change. Ideally, alignments should be made at the operating temperature after initial cold alignment of the pumping system, thus eliminating any alignment changes because of thermal expansion.

Piping. Next to the selection of operating speeds, proper piping design is the most important consideration in pump-installation design. Poor piping is often the result of inattention to details, which can lead to more than average down time, higher maintenance costs, and loss of operating-personnel confidence.

Suction piping should be direct, free of bends, as short as possible, and at least one nominal pipe size larger than the pump-suction connection. Directional piping changes should be made with long-radius elbows. A full opening block valve should be installed in the suction piping. The suction vessel should have sufficient retention time for the evolution of free gas and should be equipped with a vortex breaker at the discharge nozzle. The suction and bypass lines should enter the vessel below the minimum liquid level.

Suction piping should be large enough so that the velocity limits are not exceeded. Eccentric reducers with the flat side up should be used instead of concentric reducers. Suction piping should include a suction strainer and a pulsation dampener. Suction strainers should not be installed unless regular maintenance can be assured. A fluid-starved condition resulting from a plugged strainer can cause more damage to the pump than solids ingestion.

The discharge piping should be direct, free of excessive bends, and at least one nominal pipe size larger than the pump-discharge connection. Directional piping changes should be made with long-radius elbows. Concentric reducers may be used, but they should be placed as near to the pump as practical. To facilitate priming and starting, a bypass (recycle) line with check valve and block valve should be installed to the suction source. If a pulsation dampener is not included in the initial installation, a flanged connection should be provided if pulsation attenuation may be required. A relief valve should be installed upstream of the discharge block valve, in case overpressurization in the discharge piping occurs.

Pulsation Considerations. Flow from a reciprocating pump is not uniform. The oscillating motion of the plungers creates disturbances (pulsations) that travel at the speed of sound from the pump cylinder to the piping system. Pulsations are a function of the pump’s piston/plunger velocity, internal valves, and operating speed. Pulsations cause the pressure level of the system to fluctuate with respect to time.

Suction pulsations can cause the pressure level to drop instantaneously below the fluid vapor pressure, which results in cavitation. Caviation can cause failure of pump parts such as valves, crossheads, rods, etc. and high piping vibrations and failures of vents, drains, and gauge lines. Normal pipe clamps and supports may not be effective in controlling these vibrations.

Pulsations can be amplified by the acoustical resonances of the piping system, which results in pump fluid-end failures and piping failures because of the shaking caused by pressure pulsation. For simple piping layouts and low-to-moderate pump speeds, pulsation dampeners are used to attenuate the effects of pulsating flows. Pulsation dampeners are normally installed on both the suction and discharge. Dampeners can be liquid-filled; gas-cushioned, or tuned acoustical filters. For complicated and multiple-pump piping and high pump speeds, acoustical filters are used.

The design of a pulsation-dampening system is beyond the scope of this chapter. Special expertise is needed for analyzing and controlling pulsations in multipump installations.

Pump Drivers


Pump drivers include electric motors, steam turbines, expansion turbines, gas turbines, and internal-combustion engines.

Electric Motors

Three-phase alternating-current induction motors are the most commonly used driver for pumps because of the desirable characteristics of electricity as a power source and because the standard rotative speeds (1,750 and 3,500 rev/min) are well suited for driver centrifugal pumps. See the chapter on Electrical Systems in this volume of the Handbook for further discussion.

Steam Turbines

Large gas plants containing boilers use steam turbines to drive large pumps such as lean-oil pumps, boiler feed-water pumps, and solvent-circulation pumps. It is a common practice to select a turbine rated at pump speed and power requirements and to rely on the inherent flexibility of the turbine to provide for a margin of error.

Expansion and Hydraulic Turbines

High-pressure process streams in gas plants commonly have pressure reduced for further processing. This energy can be recovered by expansion turbines, which in turn drive pumps or generators. If the stream to be expanded is liquid (i.e., rich oil or solvents), the type of turbine is generally a centrifugal pump with the inlet and outlet reversed and is called a hydraulic turbine. If the stream to be expanded is gas, the type of turbine is called an expander.

High-pressure liquid enters the hydraulic turbine through the centrifugal pump’s discharge flange and the low-pressure fluid exits through the pump’s suction flange. Hydraulic turbines have proved economical when the developed power exceeds 160 BHP, the flow exceeds 200 gal/min, and the inlet pressure exceeds 200 psig.

The most important design consideration is handling flow and pressure-differential fluctuations so that adequate shaft power can be provided to the driven equipment at all times. This is done in one of two ways: a supplemental electric motor can be provided in tandem with the hydraulic turbine, or a continuous bypass can be installed around the hydraulic turbine that absorbs all process flow variation, always leaving the minimum required flow through the turbine.

A critical factor in the design of the hydraulic turbine is the rate of gas evolution as it flows through the turbine. Design specifications should include a complete analysis of the liquid stream so that the designer can optimize the flow passages. A pump failure by overspeed exists when the flow through the driver pump is suddenly reduced or stopped before flow through the turbine is shut off. To avoid this type of failure, an overspeed trip device should be specified.

Hydraulic turbines normally use mechanical shaft seals identical to those applied to pumps. External flushing is generally required to either prevent gas evolution within the mechanical seal area or to minimize dirt and other solids from entering this seal area.

Gas Turbines

The trend toward plants with minimal operator attendance indicates that steam may become obsolete. Gas-fired turbines are the logical choice to replace the steam turbine for large pump drives. See the chapter on Prime Movers in this volume of the Handbook for further details.

Internal-Combustion Engines

Internal-combustion engines are used to drive pumps when other power is not available or when a standby energy source is desirable. The most common application in upstream production operations is for fire water pumps and for crude-oil shipping pumps in offshore applications. See the chapter on Prime Movers in this volume of the Handbook for further details.


a = piston-rod cross-sectional area, L2, in.2
A = plunger or piston area, L2, in.2
d = inside pipe diameter, L, in.
D1 = old diameter, L, in.
D2 = new diameter, L, in.
e = the pump efficiency factor obtained from the pump manufacturer, fraction
em = motor efficiency, fraction
E = electrical energy, mL2/t2, kW-hour
g = acceleration of gravity (32.17 ft/sec2 at sea level), L/t2, ft/sec2
h = height of the fluid column above a reference point, L, ft
hf = friction loss between Points 1 and 2, L, ft
HA = net positive suction head available, dimensionless
Hd = discharge head of liquid being pumped, L, ft
Hs = suction head of liquid being pumped, L, ft
Htd = total dynamic head required of a pump, L, ft
Htd = TDH per stage at the BEP, ft
HV = the head required to overcome friction as a result of flow, L, ft
Hws = the static difference between the suction and the discharge at zero flow, L, ft
H1 = height of liquid suction vessel above pump reference point, L, ft
H2 = operating or normal height of liquid in the discharge vessel above the pump reference, L, ft
LS = stroke length, L, in.
m = number of pistons or plungers
N = pump rotative speed, rev/min
Ns = pump specific speed, dimensionless
p = pressure, m/Lt2, psi
pA = atmospheric pressure, m/Lt2, psia
pf1 = pressure drop resulting from friction in the suction piping, m/Lt2, psi
pf2 = pressure drop resulting from friction in the discharge piping, m/Lt2, psi
pva = liquid vapor pressure at pumping temperature, m/Lt2, psia
p1 = suction-vessel operating pressure, m/Lt2, psig
p2 = discharge-vessel operating pressure, m/Lt2, psig
PB = brake horsepower, horsepower
Pc = discharge flow-control-valve losses, psi
PH = hydraulic horsepower, horsepower
q = pump capacity, L2/t, gal/min
Q = flow rate, L2/t, gal/min
s = cylinder displacement, gal/min
S = slip, percent
t = time, t, days
v = average velocity of the liquid in the pipe, L/t, ft/sec
Z = height above a datum, L, ft
γ = specific gravity of the liquid, dimensionless
ρ = density, m/L3, lbm/ft3
ρf = density of the liquid being pumped, m/L3, lbm/ft3
ρw = density of water at standard conditions of temperature and pressure (62.34 lbm/ft3), m/L3, lbm/ft3


1,2 = locations along a pipe


  1. 1.0 1.1 1.2 1.3 1.4 Hydraulic Institute Engineering Data Book, second edition. 1991. Parsippany, New Jersey: Hydraulic Inst.

SI Metric Conversion Factors

ft × 3.048* E − 01 = m
ft3 × 2.831 685 E − 02 = m3
°F (°F − 32)/1.8 = °C
gal × 3.785 412 E – 03 = m3
hp × 7.460 43 E − 01 = kW
in. × 2.54* E + 00 = cm
lbm × 4.535 924 E − 01 = kg
psi × 6.894 757 E + 00 = kPa


Conversion factor is exact.