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Foot valves are specially designed check valves used at the inlet of the suction lift line to maintain pump prime, by maintaining liquid over the first-stage impeller (see Figure E.1). Foot valves are designed to open with very little pressure differential across the valve. They should be installed in a vertical orientation (or they may not work), below the top of the waterline, and the end of the inlet (suction) line should be at least four pipe diameters below the top of the water level. This will maintain a primed condition in the inlet line. The foot valve and pipe should be sized to minimize inlet line losses that will maximize the NPSH available to the pump. Problems with leakage and failure to close may be encountered where solids are present in the liquid. Therefore, foot valves may be limited to pump installations where pump nonperformance due to foot valve failure does not place the user at high risk. It is important to note that unless a suction pressure relief is fitted, pumps must have the suction side designed to contain the maximum allowable working pressure of the pump, plus any hydraulic shock loading or water hammer due to sudden foot valve closing. In vertical pumps, foot valves may be used at the inlet of the bowl assemblies for well pumps to keep the column pipe filled, to prevent backspin, and prevent well disturbance from rapidly draining water. This practice is very limited and occasionally used on small, less than 5-hp (3.7-kW) pumps, with less than 100-ft (30-m) settings and 50-psig (345-kPa) surface pressure. The user is also encouraged to check with the vertical turbine manufacturer for warranty ramifications when using a foot valve. For further information about foot valve applications in pump and piping systems, see ANSI/HI 9.6.6 Rotodynamic Pumps for Pump Piping.
The net axial downthrust force is carried by the pump shaft. The shaft will stretch, i.e., elongate, under this load. Before the pump starts up, any stretch that occurs is due to rotor weight, the sum of the static forces. The thrust load will increase after the pump starts up due to the addition of the dynamic forces.
The dynamic forces creating thrust on a vertical turbine pump enclosed impeller (Figure 220.127.116.11.3a) are due to the difference in pressure distributions on the upper and lower shrouds along with the force from the change in momentum of the flow through the impeller.
The semi-open impeller (Figure 18.104.22.168.3b) has only an upper shroud. The difference in pressure distributions along both the backside and the vaned side of the shroud is typically greater than between upper and lower shrouds of an enclosed impeller. Semi-open impeller axial thrust is higher than that of the enclosed impeller.
The axial flow pump impeller (propeller) has no upper or lower shroud; vanes are attached directly to the hub. The axial thrust generated is primarily from dynamic forces created by interaction of the propeller vanes with liquid.
The impeller back ring with balance holes configuration (Figure 22.214.171.124.3c) reduces the axial thrust. Back rings may be cast integrally into impellers with a top shroud. They are used when pump total axial thrust requires reduction. The flow through balance holes in the impeller hub shroud, combined with the leakage past the balance ring, reduces efficiency. The exact efficiency reduction depends on the individual design and pump size and specific speed. The effect of increased leakage through clearances due to wear of the back ring arrangement may be an increase in downthrust and should be considered in sizing the thrust bearing.
For more information about axial thrust for vertical rotodynamic pumps, see ANSI/HI 2.3 Rotodynamic (Vertical) Pumps for Design and Application.
There are four typical arrangements for determining the NPSH3 characteristics of rotodynamic submersible pumps. For all arrangements, the flow towards the pump must be uniform and free of undue disturbances. A pump tested with suction piping may require a flow-straightening device before entering the pump. Arrangements for cooling or heating the liquid in the loop may be needed to maintain the required temperature.
In one such arrangement, see Figure 126.96.36.199c, the pump is supplied from a closed tank in which the level is held constant. The Net Positive Suction Head Available (NPSHA) is adjusted by either varying the air or gas pressure over the liquid, varying the temperature of the liquid, or both. This arrangement tends to strip the liquid of dissolved air or gas. It is also acceptable to test with a closed loop without the closed tank on the suction side.
In another arrangement, see Figure 188.8.131.52d, the entire pump is mounted in an enclosed tank to allow the NPSH testing to be done without the suction piping connection. The testing for this arrangement is normally done at a constant flow rate while varying the NPSHA by adjusting the air pressure over the liquid in the suction tank.
In each arrangement, water must be used as the test liquid. Taking the following precautions will minimize aeration: a) No cascading return flow outlets. b) Reservoir sized for long retention time to allow air to escape. c) Inlet line properly located to prevent vortexing. d) Reservoir baffles to isolate inlet from return line. e) Tight pipe joints to guard against air leakage into the system. For more information about NPSH tests for rotodynamic submersible pumps, see ANSI/HI 11.6 Rotodynamic Submersible Pumps for Hydraulic Performance, Hydrostatic Pressure, Mechanical, and Electrical Acceptance Tests.
The necessary tools used for checking the alignment of a flexible coupling are a straightedge and a taper gauge or a set of feeler gauges. The faces of the coupling halves should be spaced far enough apart so that they cannot strike each other when the driver rotor is moved axially toward the pump as far as it will go. A minimum dimension for the separation of the coupling halves and misalignment limits are specified by the manufacturer.
Proceed with checks for angular and parallel alignment by the following method only if the face and outside diameters of the coupling halves are square and concentric with the coupling bores. A check for angular alignment is made by inserting the taper gauge or feelers between the coupling faces at 90° intervals (see Figure A.4).
The unit will be in angular alignment when the measurements show that the coupling faces are the same distance apart at all points. A check for parallel alignment is made by placing a straightedge across both coupling rims at the top, bottom, and at both sides. The unit will be in parallel alignment when the straightedge rests evenly across both coupling rims at all positions (see Figure A.5).
Allowance may be necessary for coupling halves that are not of the same outside diameter. Angular and parallel misalignments are corrected by means of shims under the motor mounting feet. After each change, it is necessary to recheck the alignment. Adjustment in one direction may disturb adjustments already made in another direction. The user is encouraged to start with shims under all motor feet so it can be raised or lowered during initial or subsequent aligning procedures. When units are aligned cold, it may be necessary to make allowance for the vertical rise of the driver and/or pump caused by heating or to verify alignment at operating temperature. Refer to instructions supplied by pump manufacturer for specific alignment requirements. Other popular alignment methods include dial indicator alignment and laser alignment. Procedures for spacer couplings, limited end float couplings, and gear type couplings are similar but may require additional steps.
For more information about coupling alignment for rotodynamic pumps, including additional methods, see ANSI/HI 1.4 Rotodynamic (Centrifugal) Pumps for Manuals Describing Installation, Operation, and Maintenance.
A sudden power and check valve failure during pump operation against a static head will result in reverse pump rotation. If the pump is driven by a prime mover offering little resistance while running backwards, the reverse speed may approach its maximum consistent with zero torque. This speed is called reverse runaway speed. If the head, under which such operation may occur, is equal to or greater than that developed by the pump at its best efficiency point during normal operation, then the runaway speed may exceed that corresponding to normal pump operation. This excess speed may impose high mechanical stresses on the rotating parts both of the pump and the prime mover and, therefore, knowledge of this speed is essential to safeguard the equipment from possible damage.It has been found practical to express the runaway speed as a percentage of that during normal operation. The head consistent with the runaway speed in this case is assumed to be equal to that developed by the pump at the best efficiency point.
The ratio of runaway speed (nro) to normal speed (nno) for single and double suction pumps varies with specific speed. This relationship is shown by Figures A.7 and A.8. The data shown should be used as a guide, because it is recognized that variations can exist with individual designs.
Transient conditions, during which runaway speed may take place, often result in considerable head variations due to surging in the pressure line. Because most pumping units have relatively little inertia, surging can cause rapid speed fluctuations. The runaway speed may, in such a case, be consistent with the highest head resulting from surging. Therefore, knowledge of the surging characteristic of the pipeline is essential for determining the runaway speed, and this is particularly important in case of long lines. For further review on reverse runaway speed for rotodynamic pumps, see ANSI/HI 2.4 Rotodynamic (Vertical) Pumps for Manuals Describing Installation, Operation, and Maintenance.
Some centrifugal pump head – rate-of-flow curves exhibit a characteristic commonly referred to as droop (see Figure 184.108.40.206.8a). A drooping head – rate-of-flow curve is one where the zero rate of flow head (shutoff head) is lower than the maximum head on the curve. This phenomenon often occurs in low to medium specific speed pumps, ns < 68 (Ns < 3500) that have been designed to optimize efficiency. Droop does not present an application problem unless one or more of the following conditions exist:
Baseplate designs for rotodynamic pumps include the grout type and nongrout type, among others.
The grouted baseplate is designed to allow grout to be poured underneath the base. The grout placed inside the base contributes to the baseplate’s installed rigidity and damping. See Figure 220.127.116.11.1a.
The cross members used on this type of base are normally designed to lock into the grout and further resist deflection or vibration of the baseplate. Typically the cross member geometry chosen to achieve this is an L-section (shown), a T-section, or an I-section. If the baseplate is a closed design (i.e., grout cannot be poured inside the baseplate perimeter due to the presence of a drain pan or deck plate), then grout holes must be provided to allow the grout to be placed inside the base. The grout used may be either cementitious or epoxy-based. The surface preparation required for a baseplate to successfully bond to the grout is different depending on which grout will be used. It is therefore important that the vendor and customer agree in advance as to which type of grout will be used. The baseplate described and shown in Figure 18.104.22.168.1a is typical of a fabricated baseplate. Cast-iron baseplates are another type of grouted baseplate. The ability to integrally cast in features such as bracing, grout holes, and sloping surfaces provides a highly functional and economical solution for many applications.The nongrout type of baseplate is placed directly on a foundation without the use of grout to fill the interior of the base to lock it to the foundation. Because of the loss of stiffening normally provided by the grout, nongrout bases must typically be structurally stiffer than comparable grouted bases. Cross members do not need to lock into the grout and so may be selected on the basis of providing the best stiffening effect. For this reason hollow rectangular sections are often used on this design. The advantage of this design over a grouted design is the simplified installation requirements. The structural design must be stiffer than an equivalent grout-type base. Additionally the structural natural frequencies must be separated from any equipment operating frequencies. This is because the stiffening and dampening effects of the grout are not available. The nongrout type is often used in offshore installations where the mass of grout and concrete must be avoided. This type can be provided for all types of rotodynamic pumps. (A multistage, axial split, between-bearings pump is illustrated.)
For more information about baseplate designs, see ANSI/HI - 1.3 - Rotodynamic Centrifugal Pumps for Design and Application
The net positive suction head (NPSH) requirements of rotodynamic vertical pumps are normally determined on the basis of handling water at or near normal room temperatures. Operating experience in the field has indicated, and a limited number of carefully controlled laboratory tests have confirmed, that pumps handling certain hydrocarbon liquids, or water at temperatures significantly higher than room temperature, will operate satisfactorily with less net positive suction head available (NPSHA) than would be required for cold water (20 °C [68 °F]). The consistency of results obtained on tests conducted with both water and hydrocarbon liquids suggests that the net positive suction head required (NPSHR) of a rotodynamic vertical pump may be reduced when handling any liquid having relatively high vapor pressure at pumping temperature. However, available data are limited to the liquids for which temperature and vapor pressure relationships are shown on Figures 22.214.171.124b, thus application of these charts to liquids other than hydrocarbons and water is not recommended except where it is on an experimental basis.
Until specific experience has been gained with operation of pumps under conditions where Figure 126.96.36.199b apply, NPSH reduction should be limited to 50% of the NPSH required by the pump for cold water. When using these charts for high-temperature liquids, and particularly with water, due consideration must be given to the susceptibility of the suction system to transient changes in temperature and absolute pressure, which might necessitate provision of a margin of safety of NPSH far exceeding the reduction otherwise available for steady state operation. Because of the absence of available data demonstrating NPSH reduction greater than 3 m (10 ft), the chart has been limited to that extent and extrapolation beyond that limit is not recommended.Figure 188.8.131.52b is based on pumps handling liquids without entrained or dissolved gas. When entrained or dissolved gases are present in a liquid, pump performance may be adversely affected, even with normal NPSH available, and would suffer further with reduction in NPSH available. Where dissolved gases are present, and where the absolute pressure at the pump inlet would be low enough to release such noncondensables from solution, the NPSHA may have to be increased above that required for cold water to avoid deterioration of pump performance due to such release. For hydrocarbon mixtures, vapor pressure may vary significantly with temperature, and specific vapor pressure determinations should be made for actual pumping temperatures. To learn more about this topic, see ANSI/HI 2.3 Rotodynamic Vertical Pumps of Radial, Mixed, and Axial Flow Types for Design and Application.
The function of a discharge casing is to collect output from the rotating impeller, decrease the velocity momentum of liquid leaving the impeller before it reaches the next stage impeller or pump discharge, and to transform increased kinetic energy of liquid at the impeller outlet into pressure.
The single volute is the most common casing style due to relative ease of manufacture and accessibility for inspection. An impeller discharges into a single spiral-shaped passage with one cutwater (tongue) that directs the liquid into the system or into the next stage of a multistage pump. The volute has a constantly increasing area cross section from the tongue, around the casing, to the discharge nozzle. The typical design criterion (see Figure 184.108.40.206) is for the liquid exiting the impeller to maintain either a constant mean velocity, constant velocity momentum or slow slightly through the spiral to the discharge at the design point.Radial thrust on the impeller varies with pump rate of flow, being lowest near the best efficiency point (BEP), and higher at reduced or increased flow rates. The thrust at off BEP can be very high for large-diameter impellers producing high head. Radial thrust also varies with impeller diameter, impeller width, and total head. Shaft deflection, combined maximum stress and bearing loads must be kept within acceptable limits by various means for best operation.
With the double volute casing, an impeller discharges into two spiral passages with two cutwaters (tongues). The pumped fluid then discharges via these two passages into a system or into the next stage of a multistage pump. The cutwaters are usually diametrically opposed in the casing. Care must be taken in the design to minimize the loss of pump efficiency. With properly designed passages, radial thrust is minimized, especially at off BEP flows. The double volute design (see Figure 220.127.116.11) is typically used to reduce shaft deflections and bearing loads to permit use of a smaller shaft and bearing sizes or to prolong the life of the pump. The casing complexity is greater than that for the single volute type due to the inaccessibility of the outside chamber.
For more information about casings for rotodynamic pumps, see ANSI/HI 1.3 Rotodynamic Centrifugal Pumps for Design and Application.
The two most common packed stuffing box configurations for vertical pumps are those with and without lantern rings. Both arrangements have a bushing below the packing. The two figures show these two constructions:The construction in Figure 18.104.22.168.4a is used where the pump discharge pressure is not high, and where pumped fluid is clean and its leakage to atmosphere is acceptable.
The type of stuffing box in Figure 22.214.171.124.4b is used to provide water injection to the shaft-enclosing tube (inner column). A provision for grease injection to the lantern ring near the lower end of the packing ring stack is optional.
Other special-purpose packed stuffing boxes provide for cooling, throat bushings, quench glands, and other special features for the particular application conditions. Packed stuffing boxes normally are limited to moderate pressures and temperatures and require a slight leakage for packing lubrication and cooling. Care is required in adjusting the packing gland to avoid shaft sleeve and packing damage. The number of packing rings in the stuffing box, together with the size and type of packing vary by manufacturer. In most cases, it is recommended that specifications leave open the exact details about the number of rings or the size or type of packing and allow the pump manufacturer to make recommendations based on application experience.
For additional information on stuffing box configurations and vertical pumps, see ANSI/HI - 2.3 - ROTODYNAMIC VERTICAL PUMPS OF RADIAL, MIXED, AND AXIAL FLOW TYPES FOR DESIGN AND APPLICATION
By Hydraulic Institute definition, the required NPSH of a pump is the NPSH available that will cause the total head (first-stage head of multistage pumps) to be reduced by 3%, due to flow blockage from cavitation vapor in between the impeller vanes. The required NPSH qualified by this criterion is referred to as NPSH3. The full published pump head will not, however, be achieved (by definition) when the NPSHA equals the NPSH3 of the pump (see Figure 126.96.36.199b). The value of 3% head drop for NPSH3 is based on accepted industry practice for defining a condition of head breakdown due to cavitation.Margin above NPSH3 is necessary in order for the pump to develop its full-published performance. Most pumps can operate satisfactorily with minimal margin above the NPSH3 value when operating near the best efficiency point (BEP) rate of flow. A greater margin may be required when operating throughout a wider operating region.
For more information about NPSH, see ANSI/HI 9.6.1 Rotodynamic Pumps – Guideline for NPSH Margin.
Either a pyrometer, thermometer, or thermocouple, should be placed on the outer surface perpendicular to the shaft centerline, over the center of the bearing(s) being recorded (see Figure A.11). On pumps with horizontal shafts, the instrument should be placed as close as possible to a vertical position. The instrument should be placed between structural ribbing when ribbing is part of the design.The pump should be operated at rated conditions. When there are differences in specific gravity or viscosity between test and jobsite liquid, adjustment to test bearing temperatures must be agreed to by all parties prior to testing. Cooling plans should be installed and be operational if necessary to duplicate field conditions. This should be agreed to by all parties.
Temperature readings shall be taken every 10 minutes for the first hour and every 15 minutes until stabilization. (Basic temperature stabilization usually occurs after the first 45 minutes. However, some bearings take up to 24 hours to stabilize and should be noted by all parties before the start of the test.) Stabilization is defined as three consecutively recorded readings taken over intervals of at least 15 minutes that fall within a 2 °C (3.6 °F) band when adjusted for a change in ambient temperature, if it occurs. When testing with a TEFC motor, the air flow from the motor should be blocked from the bearing housing where testing is being conducted. Tests have shown that the motor air flow can cause as much as an 11 °C (20 °F) false temperature reading. Similarly, the ambient air must be still. Circulating fans and opened windows can cause false readings.
For more information on other troubleshooting topics, see ANSI/HI 1.4 Rotodynamic (Centrifugal) Pumps for Manuals Describing Installation, Operation, and Maintenance.
Yes, many rotodynamic vertical pump types may be adapted to operate in reverse rotation as hydraulic turbines.While running in the turbine mode, the performance characteristics of a pump as turbine (PAT) differ significantly from pump operation. (See Figure A.5.) The applied head is usually constant, so the other parameters are shown as they vary with speed. The discharge nozzle of the pump becomes the inlet of the turbine, the suction nozzle of the pump becomes the outlet of the turbine, and the impeller of the pump, rotating in reverse direction, becomes the runner of the turbine. The impeller orientation to the casing is the same for both pump and turbine.Fixed-geometry reverse running pumps are an alternative to custom-designed adjustable gate turbines. The efficiency of a pump operating as a turbine is comparable to the pump efficiency. Many vertical pumps, with some modification, are suitable and capable of operating as turbines. Because of the reverse rotation, it must be ensured that the bearing lubrication system will operate in reverse, and threaded shaft components, such as impeller locking devices, cannot loosen. The power output is the rotational energy extracted by the reverse running pump. Its value is calculated in a similar manner as for a pump except for the placement of the efficiency term. Special care should be taken in PAT applications to ensure that the mechanical design of the unit will allow safe operation. Frequently these applications subject the pump as turbine to increased mechanical stresses, torque, and speed levels beyond original pump design values. Additionally, the turbine characteristics are such that both hydraulic forces and torsional stresses increase with increasing rate of flow. All pumps applied as turbines should be subjected to a careful calculation of combined stresses in shafts. Pumps operated in reverse as turbines tend to have relatively narrow operating bands, compared to variable nozzle (adjustable gate) turbines. At constant speed, the power developed and efficiency drop to zero at approximately 40% of the hydraulic turbine best efficiency discharge rate of flow (see Figure A.6). Energy must be added to the hydraulic turbine for it to rotate at the constant speed below this rate of flow. Changing the runner (impeller) diameter has little effect on adjusting the performance of a hydraulic recovery turbine. These facts, coupled with the difficulty in predicting hydraulic turbine performance from pump performance, result in some uncertainty when applying a pump to a power recovery turbine application unless actual test data are available on the specific pump or model pump running in reverse as a turbine. Precautions should also be taken to ensure that the pump as turbine operates without excessive cavitation. The pump industry uses the term exhaust head in place of suction head in reference to the backpressure energy level required at the outlet of the runner. There is no established industry definition for backpressure energy in vertical pump as turbine applications; however, for practical purposes the calculation is the same as for NPSH in pumps.
For additional information on rotodynamic vertical pumps, see ANSI/HI - 2.3 - ROTODYNAMIC VERTICAL PUMPS OF RADIAL, MIXED, AND AXIAL FLOW TYPES FOR DESIGN AND APPLICATION
These pumps can be quite large in capacity and dimension. Flow rates can exceed 200,000 US gpm (45,000 m3/h), heads can go up to 300 ft (90 m), and discharge nozzle sizes from 3 in (75 mm) to 72 in (1800 mm).Dry pit pumps are popular due to the ease of maintenance, and easy diagnosis of pumping problems. For sewage applications, the pump can be serviced without pulling the pump from a pit and thus less decontamination is required. Comminution1 and/or adequate bar screens must be provided to prevent large solids from entering the pump. When used, bar screen openings should be sized to prevent clogging from irregularly-shaped solids. For sewage service, pumps built to the manufacturer’s material specifications are ordinarily used.
“Nonclog pumps” are rotodynamic centrifugal pumps that are designed to ensure maximum freedom from clogging when handling liquids containing solids or stringy materials. These pumps are often used in wastewater services.
Nonclog pumps are also referred to as dry pit pumps and are recommended for handling raw or unsettled sewage, activated sludge, industrial wastewaters containing solids, and similar liquids where excessive clogging would otherwise be encountered. Dry pit pumps come in a wide range of sizes and may utilize standard motors. They are available in horizontal or vertical configurations for space-saving considerations. Pump types OH1, OH3A, and OH5A are generally applicable. These are single stage, overhung, horizontal foot or frame mounted; vertical frame mounted, including close coupled, flexibly coupled, or independently mounted motor with extended universal joint shaft drive. (See Figures A.6 – A.8.) Nonclogging impellers are commonly used in all configurations.
Corrosion-resistant and wear-resistant shaft sleeves and wearing rings are desirable for maximum life. Inspection openings in the casing or adjacent piping, for access to the impeller, are recommended. Stuffing boxes may be furnished with mechanical seals or packing, either water or grease lubricated. When water is used for the stuffing box or wearing ring lubricant or flush, the supply line must be isolated from any potable water system. For information on other application for rotodynamic centrifugal pumps, see ANSI/HI 1.3 Rotodynamic Centrifugal Pumps for Design and Application.
Vibration problems can be caused by hydraulic resonance of the liquid within the pump/piping system. Hydraulic resonance occurs when a piping system has a hydraulic resonant frequency that is excited by forces induced by operation of a pump. When normal pump-induced pressure pulsations are reflected by the piping systems and added in phase to the source pulse, the amplitude of the pulse is magnified. Resulting high pressures can ultimately cause mechanical fatigue failures in either the piping or the pump components, generate unacceptable noise levels, and cause vibration of the system components.
It is important to realize that a certain level of pressure fluctuation is unavoidable and has no detrimental effect. Excessive pressure pulsations, however, can excite pump and pipe vibrations and might even cause damage. Some typical problems include:
Experience has shown that the following measures may prove effective in correcting hydraulic resonance:a) Alter the resonant piping.b) Change the pump speed.c) Change the internal design characteristics of the pump.d) Insert a pulsation damper on the pump/piping system.For information on rotodynamic pump vibration, see ANSI/HI 9.6.4 Rotodynamic Pumps for Vibration Measurement and Allowable Values.
Inlet flow disturbances, such as swirl, unbalance in the distribution of velocities and pressures, and sudden variations in velocity can be harmful to the hydraulic performance of a pump, its mechanical behavior, and its reliability. Usually the higher the energy level and specific speed of a pump, and the lower the NPSH margin, the more sensitive the pump’s performance is to suction conditions.
All inlet (suction) fitting joints shall be tight, especially when the pressure in the piping is below atmospheric, to preclude air leaking into the fluid. Any valves in the inlet (suction) line should be installed with stems horizontal to eliminate the possibility of air accumulation. For pumps operating with a suction lift, the inlet (suction) line should slope constantly upwards toward the pump, with a minimum slope of 1% (see Figure 188.8.131.52). For most pumping systems, an inlet (suction) shut-off valve should be installed in the suction piping for system isolation.In general, as liquid travels through a piping network, entrained air tends to rise to the highest point. If the pipeline slopes upward, then the velocity of the liquid will move the air bubbles towards this high point. In contrast, if the pipeline is fairly flat and the inside surface of the pipe is very rough, or the pipeline slopes downward, the fluid velocity may not be sufficient to keep the air bubbles moving. As a consequence, it is possible for a pocket of air to collect at high points and gradually reduce the effective liquid flow area. This reduction in area can create a throttling effect similar to a partially closed valve. It is also possible that a slug of air may be swept into the pump during a restart, causing a partial or complete loss of pump prime, especially where the inlet (suction) line is kept full by a foot valve at its intake. Any amount of entrained gas in the fluid may adversely affect pump performance. Check with the pump manufacturer to determine allowable levels of entrained gas. For information on other piping requirements, see ANSI/HI 9.6.6 Rotodynamic Pumps for Pump Piping.
It can be seen from Figure A.4 that even with small percentages of gas, the unit stops pumping liquid due to accumulated gas in the impeller when operating near the shut-off condition of the pump. High velocities can carry higher percentages of gas. Therefore, when gas entrainment is a potential problem, pumps should be operated at or beyond the BEP rate of flow specified by the manufacturer. The most dramatic effect of gas or vapor on rotodynamic pump performance is the blocking of the impeller inlet. When this happens, the pump becomes airbound and the impeller acts as a centrifuge, tending to separate the heavier liquid from the gas that builds up at the impeller inlet. At low rates of flow, the liquid flow cannot even carry the gas through the impeller, and the gas bubble grows until it completely fills the impeller eye (suction side). The result is complete cessation of liquid flow.
Even when small amounts of gas are carried through the impeller, the liquid rate of flow and pump head are reduced. (See Figure A.4). This reduction is the result of the blockage of the flow by the gas, and a reduction in developed head due to the reduced specific gravity of the pumped mixture. Dissolved gases evolving out of solution can have similar effects. When the specific gravity of liquid alone is used to convert pressure to head, a lower head measurement is indicated. Laboratory tests have shown that pumps with higher specific speed (above 60 ) are affected less by the presence of gas than those with low specific speed (below 20 ). The trend of gas affecting pump performance as a function of specific speed is only approximate and specific pump designs vary widely in their gas handling characteristics. In some cases, it may be helpful to use an inducer or higher specific speed impeller in the first (suction) stage of the pump. For further discussion on this topic, see ANSI/HI 2.3 Rotodynamic Vertical Pumps of Radial, Mixed, and Axial Flow Types for Design and Application.
Pumps intended to operate in series must be close in size due to the rate of flow. Pumps operating in series produce head that is additive at the rate of flow at which they would run individually. Two pumps, each capable of 4400 gpm (1000 cubic meters per hour) at 165 feet (50 meters) of head, when connected in series, could deliver 4400 gpm (1000 cubic meters per hour) at 330 feet (100 meters) of head. Series operation is therefore used where higher pressures are required than the pressures that an individual pump can supply. Pumps used in series must be capable of the increased pressure levels that result from staging. For example, using lower speed booster pumps provide an increase in Net Positive Suction Head Available (NPSHA) for high-pressure pumps. See Figure 184.108.40.206.2a.Pumps for parallel operation must be close in total head developed; otherwise one or more of pumps may be forced to operate at shut off or at unacceptably low flows. Pumps operating in parallel produce a rate of flow that is additive at the head at which they would run individually. Two pumps, each capable of 2640 gpm (600 cubic meters per hour) at 115 feet (35 meters), when connected in parallel could deliver 5280 gpm (1200 cubic meters per hour) at 115 feet (35 meters) of head. See Figure 220.127.116.11.2b. In such cases the system curve will determine the final operation point. Two pumps operating in parallel will not automatically deliver twice the flow of one pump operating independently.
For more information of multiple pump performance, see ANSI/HI 1.3 Rotodynamic Centrifugal Pumps for Design and Application.
This answer provides the recommended minimum submergence of a vertical pump inlet bell to reduce the probability that strong free-surface air core vortices will occur. Submerged vortices are not believed to be related to submergence. If a submergence greater than recommended here is needed to provide the required NPSHR for the pump, the greater submergence should be used.Approach-flow skewness and the resulting circulation have a controlling influence on free surface vortices in spite of adequate submergence. The recommended minimum submergence given here is for a reasonably uniform approach flow to the pump suction bell. Highly non-uniform approach flows will require the application of vortex suppression devices.
Experimental analysis and field experience have resulted in the following empirical relationship:S = D + (0.574×Q)/(D1.5) Where: S is submergence in inches D is bell diameter in inches Q is rate of flow in gpm For more information on submergence, see ANSI/HI 9.8 Rotodynamic Pumps for Pump Intake Design.
While antifriction ball or roller bearings are used on most slurry pumps, hydrodynamic bearings may be used on some large slurry units, such as dredge pumps. Bearings may be grease or oil lubricated, and bearing housings must be effectively sealed from leakage and outside contamination. Labyrinth seals, bearing isolators, lip seals and other proprietary seals are commonly used.Contact seals include all designs that have dynamic contact as a requirement for proper function. Contact seals are recommended for applications in which the seal must retain a static level or pressure differential, such as a horizontal bearing housing with a lubricant level above the shaft seal surface.
Labyrinth seals consist of a simple gap seal with labyrinth grooves and possibly a gravity drain to augment performance. Labyrinth grooves help retain splash oil lubrication, but they rely on a simple gap for contaminant exclusion. Bearing isolators have a stationary and a rotating component that act in concert to retain lubricant and exclude contaminants from the bearing housing. Bearings should be sized for the calculated fatigue life that corresponds to the slurry service class (see Figure 18.104.22.168a) shown in Table 22.214.171.124. Slurry pumps are ranked by light (Class 1), medium (Class 2), heavy (Class 3) and very heavy (Class 4) services. The boundary lines between the service class areas in the chart approximate limits of constant wear modified for practical considerations and experience. Calculations should be done at the worst acceptable operating point, which is usually minimum flow. More severe services require a longer calculated bearing life because of the impact of large solids, possible cavitation and variable loads. Increasing the shaft and housing size may be necessary to accommodate the correct bearings for a given application. For more information on slurry pumps, see the Hydraulic Institute Standard ANSI/HI 12.1-12.6 Rotodynamic (Centrifugal) Slurry Pumps for Nomenclature, Definitions, Applications, and Operation.
The basic design requirements for satisfactory hydraulic performance of rectangular intake structures include adequate depth of flow to limit velocities in the pump bays, reduction of the potential formulation of surface vortices, and adequate pump bay width to limit the maximum pump approach velocities. The pump bay width should be narrow and long enough to channel uniform flow toward the pumps.
The minimum submergence (S) required to prevent strong air core vortices is based in part on the Froude number—a dimensionless flow parameter with consistent units—defined as:
FD=V/(gD)0.5Where:FD = Froude number (dimensionless)V = Velocity at suction inlet equals the flow per unit area, based on DD = Outside diameter of bell or pipe inletg = Gravitational accelerationThe minimum submergence with units consistent with the outside diameter of the bell or pipe inlet can be calculated using the formula:S=D(1+2.3FD)There is some variation in bell velocity among pump types and manufacturers. However, variations in bell inlet velocity are secondary in maintaining acceleration of flow and converging streamlines into the pump bell.
The effectiveness of the recommended pump bay dimensions depends on the characteristics of the flow approaching the structure and the geometry of hydraulic boundaries in the immediate vicinity of the structure.
For additional information, see ANSHI/HI 9.8 Intake Design for Rotodynamic Pumps for a basic recommended layout of rectangular sumps.
When the viscosity of the pumped liquid increases, the Reynolds number decreases causing friction factors in the hydraulic passages of the pump to increase just like flow through a pipe. Two examples of pump losses that can be anticipated are mechanical and hydraulic losses. Mechanical losses are essentially independent of the viscosity of the liquid being pumped. Hydraulic losses—similar to pipe friction losses—occur at the inlet, in the impeller, in the volute or diffuser, and in the discharge of a pump. In basic rotodynamic pump theory, the useful head is the difference of the impeller theoretical head minus the hydraulic losses. Viscosity does not generally influence the flow deflection or slip factor of the impeller. As a result, the theoretical head is not affected. Head reduction due to viscous flow is primarily a function of the hydraulic viscous flow losses. These hydraulic losses consist of friction losses, which are a function of the Reynolds number (pump size, rotor speed and viscosity effects), surface roughness of the hydraulic passageways, and mixing losses caused by the exchange of flow momentum due to non-uniform velocity distributions. Such non-uniformities or mixing losses are caused by the action of work transfer from the blades, decelerations of the liquid, angle of incidence between liquid flow and blades, and local flow separations.
For more information on the influence of a liquid’s properties in a pump, see ANSI/HI 9.6.7 Effects of Liquid Viscosity on Rotodynamic Pumps.
End users should evaluate their need for dynamic analysis by considering the level of proven field experience available for any given configuration. The vendor and user should agree on which types of analysis should be performed at any level. Lateral, torsional and/or structural analyses are three identifiable and normally separable deliverables. In all cases, it is the user's prerogative to specify additional tests, validations and/or analyses to further mitigate risk. Historically, dynamic analysis trends have developed within the various pump application markets because of the types and characteristics of equipment typically used and as a result of past experiences. In the oil and gas industry, single-stage overhung horizontal pumps and between-bearings, one- and two-stage pumps must be designed to be classically rigid, which can eliminate the need for lateral dynamic analysis . Multistage pumps identical to pumps proven in-field are also not subject to lateral analysis. Vertically suspended pumps are required to be designed with established limits on bearing spacing to ensure suitable lateral rotodynamic performance. Drive system configuration and power levels determine the need for torsional dynamic analysis. High-energy, high-speed, critical-service and unspared machines are subject to high levels of customer intervention and scrutiny, with the customer having varying definitions of these terms.
For more information on dynamic analysis, refer to ANSI/ HI 9.6.8: Rotodynamic Pumps Guideline for Dynamics of Pumping Machinery.
The performance of a rotodynamic pump on slurries will differ from the performance on water, which is the basis for most published curves. Head (H) and rate of flow (Q) will normally decrease as solids size and concentration increases. Power (P) will increase and starting torque may also be affected. This “solids effect” is shown schematically in Figure 126.96.36.199 along with the head and efficiency derating terms used.
Effects of solids on a slurry pump cavitation performance are dependent on the slurry type and the pump design and can be highly variable. The value of net positive suction head required in order not to exceed 3% head drop, NPSH3, will increase, in most circumstances.For settling slurries of low to medium concentration, a modest increase in NPSH3 can be expected. For a particular application, this increase can be conservatively estimated by dividing the value of NPSH3 on water by the head derating factor. For viscous and nonsettling slurries or slurries with entrained air, the effect on pump cavitation performance can be significantly greater. The pump manufacturer should be consulted for guidance regarding slurry effects on NPSHR.Multiple approaches can be used for predicting the rotodynamic pump performance change from water to slurry, depending on the slurry type, and pump users should consult with pump manufacturers for the most accurate predictions of performance for a particular pump and particular slurry.
For more information on rotodynamic slurry pump performance, see ANSI/HI 12.1-12.6 Rotodynamic (Centrifugal) Slurry Pumps for Nomenclature, Definitions, Applications, and Operation.
The recommended use of the NPSH margin involves a known pump design having fixed NPSH3 characteristics that result in a reasonable and safe suction specific speed. In such a situation, users should apply the NPSH margin to the NPSH3 at the flow rate of interest to obtain the minimal value of NPSHA. The use of a higher value of NPSH margin in such a situation generally results in more conservative conditions for the pump. If the recommended NPSH margin cannot be obtained, choosing a lower operating speed for the pump for a fixed flow rate will generally result in a conservative selection.
Users should be cautious when obtaining NPSH margin by specifying pumps with higher suction specific speeds that have lower NPSH3 values. Higher suction specific speed pump designs are more likely to experience objectionable noise and a narrower operating range as compared with lower suction specific speed pump designs.Extra margin may be necessary to account for changes in pump geometry that can increase NPSH3. For example, erosion can enlarge impeller running clearances and increase internal leakage at the impeller eye, adversely affecting the NPSH3.
Added NPSH margin may be needed either to cover uncertainties in the NPSHA or because of the actual operating rate of flow. If a pump consistently operates at higher than expected flow rates, then the NPSHA of the system may be lower than expected and the NPSH3 for the pump will be higher, providing a smaller—or possibly negative—NPSH margin.
Poor suction conditions may result in flow separation and distorted flow at the impeller inlet, which can adversely affect the NPSH available to the pump. The negative effects of poor inlet/suction conditions are difficult to calculate. Greater NPSH margins might not correct the problem. Optimum pump performance requires users to follow proper intake/suction piping design to ensure a steady, uniform flow to the pump suction at the required suction head.
The determination of an adequate NPSH margin is dependent on an accurate prediction of both pump NPSH3 and system NPSHA. Factory NPSH3 tests can help validate the predicted pump suction performance. Determination of the NPSHA is the responsibility of the system designer.
For more information on net positive suction head, please refer to ANSI/ HI 9.6.1 Rotodynamic Pumps Guideline for NPSH Margin.
The sealless pump is used when there is a need to contain toxic, dangerous and/or valuable fluids. Application may be dictated by space, noise, environment or safety regulations. This section outlines types, nomenclatures and components of sealless rotodynamic type pumps. Sealless pump design is founded on eliminating the dynamic shaft seal between the wetted end of a centrifugal pump and the atmosphere. This is achieved by enclosing the pump and its rotor assembly inside a pressure vessel with the pumped fluid. The pressure vessel or “primary containment” is sealed by static seals, such as gaskets or O-rings. The inner rotor assembly is driven by a rotating magnetic field that is transmitted through a containment barrier. Sealless pumps fall into two categories: magnetic driven pump (MDP) canned motor pump (CMP), as shown in Figures 188.8.131.52 and 184.108.40.206.
It is recognized that the liquid-lubricated bearing design and application considerations are essentially the same for CMP and MDP. Factors internal to the unit design, such as pressures, temperatures, flows and heat transfer characteristics within the drive section and hydraulic performance of the pump end must be understood to properly select circulation plans and assess application questions. Properly designed, applied and operated sealless pumps may offer the following advantages:• Improved safety when handling hazardous liquids• Eliminated leakage through primary containment to the environment during normal operation• Optional backup secondary containment• Eliminated loss of valuable liquids• Lower noise level (CMP designs)• Suction pressure usually does not affect the axial thrust• Reduced or eliminated periodic shaft seal replacement costSome limitations must be understood to properly apply a sealless pump such as:• Temperature of motor windings (CMP) or magnet components (MDP)• Control of bearing environment is required to provide clean non-flashing liquid• Primary containment shell is relatively thin and corrosion potential should be carefully considered• Retraining of maintenance personnel may be required• Drive-generated heat may affect the NPSH required with some circulation plans for volatile liquids• Overheating of drive section may occur with loss of flow or loss of suction• There is potential for higher repair cost if bearings fail before detectionFor more information about sealless pumps, see ANSI/HI 5.1-5.6 Sealless Rotodynamic Pumps for Nomenclature, Definitions, Application, Operation and Test.
A centrifugal seal is a dynamic seal that only operates when the pump shaft is rotating and has no seal effect when the shaft is stationary. It consists of an expeller or set of expellers located in a separate chamber behind the impeller, which is typically fitted with expelling vanes on the back shroud. When the pump is running, the centrifugal seal generates pressure Pe to equalize the pressure Pb, as shown in Figure 220.127.116.11.6, so that the pump operates without leakage.
A centrifugal (dynamic) seal needs to be combined with a backup or static seal to prevent leakage when the pump is not running. The general requirements for the backup-sealing device are that it must seal statically when the pump is shut down and it must run dry during pump operation. This can be accomplished by dry-type packing, multiple lip seals, other proprietary devices, or mechanical seals with either dry run capabilities or fitted with a separate flush.
There is a maximum allowable suction pressure Ps above which, depending on speed of rotation, a centrifugal seal will not operate properly. For this reason, centrifugal seals are not effective on the second or higher stages of multiple-pump installations, where the pumps are arranged to have the full discharge of the preceding stage applied to the suction of the following stage. If the pumps are installed at specified intervals and elevations spread out along a slurry transport line, then it is possible to use centrifugal seals on all stages. The arrangement should be such that the suction pressures on each stage are approximately equal and do not exceed 10 to 20 percent of the discharge pressure. An analysis should be made of the centrifugal seal performance, based on actual head, flow and suction pressure, so that proper operation is ensured.
Additional information about centrifugal (dynamic) seals is available in ANSI/HI 12.1-12.6 Rotodynamic (Centrifugal) Slurry Pumps for Nomenclature, Definitions, Applications, and Operation.
NPSH3 is the net positive suction head available to a pump under test at a constant rate of flow when the pump head is decreased by 3 percent as a result of cavitation caused by a decreasing available suction head. Sometimes NPSH3 is referred to as net positive suction head required (NPSHR); however, a pump’s NPSHR must be higher than the NPSH3 for the pump to operate without head reduction, and it may need to be higher than the NPSH3 for long-term reliable operation. For more information on the required margins above NPSH3, refer to ANSI/HI 9.6.1 Rotodynamic Pumps Guideline for NPSH Margin.
The NPSH is established through the datum elevation of the impeller. For all arrangements, the flow toward the pump must be uniform and free of undue disturbances. A pump tested with suction piping may require a flow-straightening device before entering the pump. Arrangements for cooling or heating the liquid in the loop may be needed to maintain the required temperature.
Multiple arrangements can be used to determine the NPSH3 characteristics of rotodynamic submersible pumps. One arrangement is shown in Figure 18.104.22.168a. The pump is supplied by a constant-level supply through a throttle valve followed by a section of pipe containing straightening vanes or a minimum of seven diameters of straight pipe to straighten flow. This arrangement decreases the turbulence produced by the throttle valve and makes possible a more accurate reading of suction pressure at the pump inlet.
This simple arrangement usually is satisfactory for NPSH greater than 3 meters (10 feet); however, the turbulence at the throttle valve tends to accelerate the release of dissolved air or gas from the liquid, which takes place as the pressure on the liquid is reduced. A test made with this arrangement usually indicates higher NPSHR than what would be expected with deaerated liquid.
A second arrangement is shown in Figure 22.214.171.124b. The pump is supplied by a sump in which the liquid level can be varied to establish the desired NPSHR. End users should be careful to prevent entrained air or vortexing as the liquid level is varied. The priming connection should be installed above the eye of the impeller, either in the discharge pipe or on the pump.
For more information on test methods for rotodynamic submersible pumps, refer to ANSI/HI 11.6 Rotodynamic Submersible Pumps for Hydraulic Performance, Hydrostatic Pressure, Mechanical, and Electrical Acceptance Tests.
The net axial downthrust force is carried by the pump shaft. The shaft will stretch, or elongate, under this load. Before the pump starts up, any stretch that occurs is the result of the sum of the static forces, such as rotor weight. The thrust load will increase after the pump starts up because of the addition of dynamic forces.
The dynamic forces creating thrust on a vertical turbine pump enclosed impeller (Figure 126.96.36.199.3a) result from the difference in pressure distributions on the upper and lower shrouds along with the force from the change in momentum of the flow through the impeller.
The semi-open impeller (Figure 188.8.131.52.3b) has only an upper shroud. The difference in pressure distributions along both the backside and the vaned side of the shroud is typically greater than between the upper and lower shrouds of an enclosed impeller. Semi-open impeller axial thrust is higher than that of the enclosed impeller.
The axial flow pump impeller (propeller) has no upper or lower shroud; vanes are attached directly to the hub. The axial thrust generated is primarily from dynamic forces created by interaction of the propeller vanes with liquid.
The impeller back ring with balance holes configuration (Figure 184.108.40.206.3c) reduces axial thrust. Back rings may be cast integrally into impellers with a top shroud. They are used when pump total axial thrust requires reduction. The flow through balance holes in the impeller hub shroud, combined with the leakage past the balance ring, reduces efficiency. The exact efficiency reduction depends on the individual design, pump size and specific speed. Increased leakage through clearances that results from wear of the back ring arrangement may cause an increase in downthrust. End users should consider these effects when sizing the thrust bearing.
For more information about axial thrust for vertical rotodynamic pumps, see ANSI/HI 2.3 Rotodynamic (Vertical) Pumps for Design and Application.
The most common factors affecting pump vibration include:
Table 1 from ANSI/HI 9.6.4 can help identify possible sources of high pump vibration.
For more information on pump vibration, reference ANSI/HI 9.6.4 Rotodynamic Pumps for Vibration Measurements and Allowable Values.
vertical turbine pumps can be installed in a wet pit or closed bottom can as
shown in Figure G.1. Well motor types are recommended for both wet-pit type
and closed-bottom can type of below-grade suction intakes in a rotodynamic
A submersible well-type motor
normally requires a minimum flow of liquid around the immersed motor to
provide adequate motor cooling. For many applications, a shroud is required
to ensure proper cooling flow around the motor. Sizing of the cooling shroud
for internal flow velocities must be referred to the pump manufacturer. The
top of the shroud must include a cover to restrict downward flow of liquid to
the pump inlet while allowing for venting air from the shroud. The confined
flow pathway provided by the motor cooling shroud is very desirable in
developing a uniform flow to the first-stage impeller.
The characteristics of the flow
approaching an intake structure are among the most critical considerations
for the designer. When determining direction and distribution of flow at the
entrance to a pump intake structure, the following must be considered:
types of single-phase motors are used throughout the commercial and
industrial world. Listed below are a few types of single-phase AC motors used
in certain pumping applications:
Split-phase: A split-phase motor is a single-phase induction motor equipped
with a main winding and an auxiliary starting winding. This type of motor has
a switch that deactivates the starting winding as the motor comes up to
speed. Split-phase motors are used in spa, jetted tub and aboveground pool
pump applications. The motors are usually rated from 1/6 horsepower (hp)
through 1.5 hp.
Capacitor-start: A capacitor-start motor is a single-phase induction motor
equipped with a main winding and an auxiliary starting winding with a series
capacitor. This type of motor has a switch that deactivates the starting
winding and capacitor as the motor comes up to speed. They are usually rated
from 1/6 to 7.5 hp. Capacitor-start motors are the most common type of
single-phase motors found on in-ground pool, irrigation and dewatering pump
Permanent-split capacitor: A permanent-split capacitor motor is a
single-phase induction motor equipped with a main and an auxiliary starting
winding with a series capacitor. The motor does not have a switch and both
the main and starting windings are always energized. The motors are typically
rated from 1/2 to 15 hp and are commonly found in dewatering and irrigation
Capacitor-start, capacitor-run: A capacitor-start, capacitor-run motor is a
single-phase induction motor equipped with a main and an auxiliary winding
with a series run capacitor. Both windings and the run capacitor are
permanently energized. In addition, this motor has a second capacitor called
a start capacitor that is deactivated by a switch as the motor comes up to
speed. This type of motor is also called a two-value capacitor motor. The
motors are usually rated between 1.5 and 15 hp and are commonly found in
dewatering and irrigation applications.
For more information on centrifugal pumps, refer to ANSI/HI 1.3 Rotodynamic
Centrifugal Pumps for Design and Application.
hammer, or hydraulic shock, is a condition that exists when a column of fluid
changes velocity quickly in a piping system. There are several causes of
water hammer such as a pump start or stop or a rapid change of valve
position. The velocity change results in a pressure wave that is above and
below the normal pressure in the piping system. These pressure waves are
called surge pressures or water hammer (when water is the fluid), and their
magnitude can be sufficient to burst or collapse piping, valves, machinery
casings and other devices.
The pressure wave’s magnitude can be
calculated with reasonable precision if the user knows the configuration of
the piping, the size of the pipes, the materials of the piping, the
properties of the fluid, and how quickly the pump and/or fluid accelerates or
decelerates. When the column of fluid in the piping is either started or
stopped, the energy of the system is transformed from velocity energy to head
or pressure energy. Because the fluid and piping material are not completely
incompressible, they will absorb a fraction of the energy.
Cast iron, for example, is a rather
brittle material and is more susceptible to failure from sudden impact of a
pressure wave that is well above its normal internal pressure and traveling
at the speed of sound. Other materials that are more ductile may absorb the
shock waves without cracking but still risk permanent deformation and ultimate
failure. The pump is not the only component that is affected by this
phenomenon; valves, sprinkler heads and pipe fittings are also at risk of
catastrophic damage. Water hammer can adversely affect pipe hangers and pump
foundations. Polyvinyl chloride (PVC) pipe and fittings are very susceptible
to damage from water hammer.
Surge analysis is necessary because
surge will occur in every pumping system. Water has mass. One cubic meter of
water at 15 C (59 F) weighs 1,000 kilograms (kg) (1 cubic foot [ft3] of water
weighs 62.4 pounds, or 1 gallon weighs 8.3 pounds at sea level). Moving water
has momentum, which is directly related to both the mass and the velocity of
the liquid. The faster the liquid is flowing, the greater its momentum. The
greater the momentum, the more damage water hammer can cause if the liquid is
suddenly stopped. Surge pressure will be maximized when the fluid is stopped
in less time than it takes for a pressure wave to travel from the equipment
that stopped the flow to the other end of the piping system and back.
Water hammer can be understood
through proper surge analysis and controlled through proper valve closure
rates (with slow-closing valves), controlled starting and stopping of pumps,
the addition of diaphragm tanks to absorb the pressure surge, and relief
valves to release the pressure.
For more information on water
hammer, refer to ANSI/HI 9.6.6 Pump Piping for Rotodynamic Pumps.
NPSHA is the total suction head available over the vapor pressure of the
liquid pumped corrected to the centerline of the impeller (or impeller inlet
vane tip datum if vertically mounted) as shown in Figure 220.127.116.11a and
measured at the inlet to the pump. An NPSH margin may be required for several
reasons related to pump performance and service life. The margin can
underline the uncertainties of what the NPSHA will be over the range of
operation, and it can provide for adequate pump reliability and performance.
When calculating NPSHA based on
suction pressure gauge measurements, note that the elevation difference
between the gauge and the datum must be added to the NPSHA. If the gauge
elevation is above the datum, it will increase the NPSHA calculation. The
opposite is true if the gauge elevation is below the datum. As noted
previously, NPSH margin is required and is calculated as shown in Equation 1.
NPSH3 is the point where a 3 percent head loss is observed at a constant flow
rate as NPSHA is reduced. This is an accepted industry practice defining a
condition of head breakdown resulting from cavitation and is used in the NPSH
NPSHA=hatm + hs - hvp
hatm = atmospheric pressure head, in m (ft)
hs = total suction head = hgs+ hgs+ zs, in m (ft)
hgs = suction gauge head, in m (ft)
hvs = suction velocity head, in m (ft)
zs = elevation from the suction gauge centerline to datum (see Figure
18.104.22.168a), in m (ft)
hvp = liquid vapor pressure head (taken at the highest sustained operating
temperature), in m (ft)
For more information on NPSH, refer to ANSI/HI 9.6.1 Rotodynamic Pumps –
Guideline for NPSH Margin.
valves are an essential component in the design of system piping. These
valves are hydro-mechanical devices that vent pockets of air as they
accumulate in a system. In general, as the liquid is being pumped, entrained
air will bubble up to the high points in the system piping. If the pipeline
slopes upward, the liquid’s velocity will move the air bubbles to stagnate at
a high point. In contrast, if the pipeline is fairly flat and the inside
surface of the pipe is very rough, or the pipeline slopes downward, the fluid
velocity may not be sufficient to keep the air bubbles moving.
As a consequence, it is possible for a pocket of air to collect at these
high points and gradually reduce the effective liquid flow area, which may
create a throttling effect similar to a partially closed valve. Additionally,
sudden compression of these air pockets can set up severe shock waves when
released with the potential for serious equipment damage. When pumping into a
pressurized system, an automatic air release valve may be required.
For vertical wet pit pumps on water service, it is usually appropriate to
use both an air release valve and a vacuum valve. The valves should be
located on the high point of the pump outlet (discharge) nozzle and between
the pump outlet (discharge) and the outlet discharge check valve. It is
important to size and adjust the air-release valve to evacuate most of the
air on start-up. A small volume of air is desirable to cushion the water as
it slams against the check valve on pump start-up. A vacuum valve is usually
desirable on vertical turbine wet pit/well pumps to release the vacuum
accumulated during pump shutdown when the water in the column pipe returns by
gravity to the wet well. If the vacuum in the column is not released to
atmospheric pressure when the pump is shut down, it may result in a damaging
water slam against the check valve.
On initial filling of piping systems or after a shutdown that results in
draining of the system, caution is required to prevent possible damage to
air-release valves. During filling of the pipeline with liquid, there can be
a sudden change in momentum forces on the air-release valves when the fluid
flowing in the pipe changes from all low-density gas to the higher-density
liquid. When this interface passes a high point with an air-release valve,
the valve can undergo the same sudden change in momentum. The resulting
forces may exceed the strength of the air-release valve housing.
For horizontal pumps that are not self-venting (trap air at the top of the
volute), a valve should be located at the highest point on the volute case.
Additionally, air-release valves may be installed on pipeline high points,
changes in grade and at periodic intervals on long horizontal runs that lack
a defined high point.
For more information on specialty piping components and applications, refer
to ANSI/HI 9.6.6 Pump Piping for Rotodynamic Pumps.
For a duplex air-operated pump, compressed air is directed into the pump air chamber via the air distribution system. The compressed air is separated from the liquid by a diaphragm. The diaphragm, in turn, applies pressure on the liquid and forces it out of the pump discharge. While this is occurring, the opposite-side air chamber is depressurized and exhausted to atmosphere and liquid is drawn into the pump suction. The cycle repeats, thus creating a constant reciprocating action, which maintains flow through the pump. (See Figure 10.4.2.1).
A simplex pump operates in much the same manner except that no liquid is discharged during the recharge stroke of the diaphragm or bellows. A spring return is typically used to return the diaphragm and provide energy for the suction stroke. Intake and discharge valves direct flow into and out of the pumping chamber. It should also be noted that simplex pumps require a controller to cycle the compressed air/gas into the air chamber of the pump and to exhaust the air out. The controller can be mechanical or electrical, such as a repeat cycle timer and solenoid valve. The speed of the pump is set by the controller, and the maximum speed is determined by the manufacturer.To determine the maximum pumping speed on a duplex pump, increase the air supply while the pump rate of flow increases. When the rate of flow no longer increases, throttle back the air supply until the pump rate of flow starts to decrease. This point is the optimum pump cycle rate achievable under these system conditions.
For more information about air-operated pumps see ANSI/HI 10.1-10.5 Air-Operated Pumps for Nomenclature, Definitions, Application, and Operation.
Application of reciprocating power pumps requires careful consideration of their starting and running torque demands. These affect the selection of driver motors, motor starters, engines, gear reducers, belts or chain drives, couplings and universal joints. The effect of such loads on an electrical distribution system requires thought, especially if the pump is large.For starting the pump with a liquid bypass, the operator manually opens a bypass valve, or a power-actuated dump valve programmed to open automatically, which bypasses the liquid during the start and the stop function. A check valve is employed in the pump discharge line, and it remains shut as long as the bypass (dump) valve remains fully or partially open (see Figure 6.47).
The liquid pressure exerted on the plungers (or pistons) is largely that caused by liquid mass and friction as it passes through the bypass valve and piping. When correctly sized, the bypass valve and piping cause low back-pressure. Only a relatively small torque is required while bypassing the liquid to a tank. With liquid bypass, the total starting torque requirement is mainly related to the mechanical inertia of the pump, couplings, gears and motor rotor. These components are heavy, and substantial starting torque may be required. All the liquid in the pump suction line and in the bypass line must be accelerated from standstill to full liquid velocities.The torque needed to accelerate the entire mechanical hydraulic inertia system depends on the inertia of all the moving parts, including the liquid, the rate of acceleration and the total system friction. The rate of acceleration is very important, and the starting torque is directly proportional to it. In other words, peak torque is inversely proportional to the time duration of acceleration.A few pumps are provided with mechanical suction valve unloader devices. These devices reduce the total starting torque requirement to that needed to accelerate the mechanical items, such as pump crankshaft, gears, couplings, etc. They stop liquid pumping action by mechanically holding open the pump’s suction valves, allowing the liquid inducted into each liquid cylinder to be delivered back into the pump suction. Because no liquid pumping occurs, the driving machinery does not need to apply a torque to accelerate the liquid. Only the torque to overcome mechanical inertia and friction is needed during the start.After the pump and driver have reached full speed, the suction valve unloading devices are retracted, and normal pumping action then commences. At this point, the driver must supply additional torque to accelerate the liquid system and also to meet the total running torque requirement caused by discharge pressure.For more information about torque requirements for reciprocating power pumps, see ANSI/HI 6.1-6.5 Reciprocating Power Pumps for Nomenclature, Definitions, Application, and Operation.
Pump input power may be determined by transmission dynamometers, torsion dynamometers, strain gauge type torque-measuring devices, calibrated motors or other sufficiently accurate measuring devices.When applicable, readings of power shall be taken at the same time that rate of flow is measured.Methods of measurement of power input to the pump fall into two general categories:Those which determine the actual power or torque delivered to the pump and are made during the test by some form of dynamometer or torque meter;Those which determine the power input to the driving element, taking into account the driver efficiency when operating under specific conditions.When pump input power is determined by transmission dynamometers, the unloaded dynamometer shall be statically checked prior to the test by measuring the load reading deflection for a given torque; and by taking the tare reading on the dynamometer scale at rated speed with the pump disconnected. After the test, the dynamometer should be rechecked to assure that no change has taken place. In the event of a change of ± 0.5% of the power at the best efficiency point (BEP), the test should be rerun. An accurate measurement of speed within ± 0.3% is essential.The use of calibrated dynamometers or motors is an acceptable method for measurement of input power to the pump.Calibration of the torsion dynamometer should be conducted with the torsion-indicating means in place. The indicator should be observed with a series of increasing loadings and then with a series of decreasing loadings. During the taking of readings with increasing loadings, the loading is at no time to be decreased; similarly, during the decreasing loadings, the loading should be based on the average of the increasing and decreasing loadings as determined by the calibration.If the difference in readings between increasing and decreasing loadings exceeds 1%, the torsion dynamometer shall be deemed unsatisfactory. Dynamometers shall not be employed for testing pumps with a maximum torque below one-quarter of the rated dynamometer torque.When strain gauge type torque measuring devices are used to measure pump input horsepower, they shall be calibrated, with their accompanying instrumentation, at regular intervals (see Figure 6.72). After the test, the readout instrumentation balance shall be rechecked to assure that no appreciable change has taken place. In the event of a change of ± 0.5% of the power at BEP, the test shall be rerun.
Calibrated electric motors are satisfactory to determine the power input to the pump shaft. The electrical input to the motor is observed, and the observations are multiplied by the motor efficiency to determine the power input to the pump shaft. Calibrated laboratory type electric meters and transformers shall be used to measure power input to all motors.For more information about this topic, see ANSI/HI 6.6 Reciprocating Pump Tests.
There is a substantial first-cost premium for sealless rotary pumps. Depending on the type, this cost is a multiple of a similar pump with a conventional mechanical seal. The expense associated with magnetic drive pumps is primarily associated with the cost of the magnets required to transmit torque.There are many applications where sealless pumps may be selected based on a life cycle cost analysis. Each application is subject to individual review, but general areas that may be considered in a life cycle cost analysis follow:a) Negative for sealless pumps:Initial capital costAdditional instrumentation, if required
Lower drive efficiency
b) Variable cost elements based on user experienceCost of maintenance differential (numerous surveys have shown mechanical seal failure to be the primary cost of pump maintenance)Cost of production loss associated with system reliabilityReduced cost of instrumentation compared to complex seal systems
c) Positive for sealless pumps:Monitoring exemption for fugitive emissions under the Clean Air ActCost of liquid disposal as a result of seal failureRisk associated with personnel or fire incident lossCost of alternate sealing systems to meet environmental emissions control requirements, such as double-seal and complianceValue of extended operating periods between unit turnaroundSpace reduction when compared to complex seal systemsCooling water costsEconomics may initially be difficult to assess until operating, maintenance, and monitoring expense experience has been assembled. Records of these considerations are valuable to the pump-type selection process.For more information about rotary sealless pumps, see ANSI/HI 4.1-4.6 Sealless, Magnetically Driven Rotary Pumps for Nomenclature, Definitions, Application, Operation, and Test.
Screw pumps are used in oil field, pipeline, refinery, marine, power generation, chemical, hydraulic systems, and general industrial applications for transfer, lubrication, injection, and hydraulics handling a wide range of fluids, such as fuel oils, lube oils and greases, asphalts, noncorrosive viscous chemicals, and high-pressure coolants.The untimed rotary screw pump is an axial-flow, multirotor, positive displacement design used in a wide range of applications in pumping clean to mildly abrasive viscous liquids. It is often a more efficient alternative than centrifugal pumps. The design may use two, three, four, or five screws. The most common configuration is the three-screw pump, which consists of a power rotor (drive screw) and two symmetrically opposed idler rotors (driven screws) that mesh within a close-fitting housing forming a succession of cavities to continuously convey fluid to the pump discharge.Untimed screw pumps are available with a double-ended flow path as illustrated in Figure 22.214.171.124c or with a single-ended flow path as shown in Figure 126.96.36.199b. Timing is accomplished through rotor geometry. In a properly applied three-screw pump, there is no rotor contact because screws are supported radially in their bores and are hydraulically balanced or free to float on a hydrodynamic film created by the pumped liquid. In other untimed screw pump configurations, the screws may be supported in product-lubricated bushings.
Units are commercially available in product families with flows to 1200 m3/h (5300 gpm) and discharge pressures to 310 bar (4500 psi). Applications cover a wide viscosity range from 2 to 220,000 cSt (33 to 1,000,000 SSU) and temperatures from below zero to 274°C (500°F). Because of the axial movement of the fluid and the compact diameter of the rotors, untimed screw pumps typically operate at motor speeds (two-, four-, and six-pole). Screw pumps operate with a minimum of noise, vibration, and fluid pulsation. Other characteristics important in many applications are their good suction capability and low shear rate. Untimed screw pumps are frequently found in installations where extended uninterrupted service life is required.For more information about rotary pumps, see ANSI/HI 3.1-3.5 Rotary Pumps for Nomenclature, Definitions, Application and Operation.
Various instrumentation and systems can be used. Here are some examples:Torque meter: The most direct method to monitor the power used is to install a torque meter with integral speed pickup between the driver and the pump. This system will directly sense the speed and torque required by the pump. Some torque meter readouts will even calculate the actual power transmitted. Under controlled operating conditions, torque measurement can sometimes be used to detect internal component rubbing and wear in rotary pumps.Power meter: This measurement is useful if the pump is driven by an electrical motor, either directly coupled to the pump or through a gearbox, belt, or hydraulic coupling. Electrical transducers are typically installed in the electrical motor starter to measure voltage, current, and phase angle. The product of these values results in the power supplied to the motor. This approach will not only monitor the power increase or decrease in the pump as parts deteriorate or drag, but also indicates if anything is happening to the general health of the electrical motor and/or the gearbox, belts, or hydraulic coupling. This type of device may not be suitable for use with motors controlled by variable frequency drives (VFDs), which alter the characteristics of the voltage and current supplied to the motor.Electrical current: Similar to a power meter, but only the motor current is monitored. The line voltage and power factor are assumed to remain constant, allowing one to calculate the power supplied to the motor. While this method monitors the condition of the pump and motor, it is susceptible to error caused by variations in the electrical supply grid and when motors are running at low load.Strain gauges: Strain gauges applied to either the pump shaft near the coupling, or the drive output shaft, with proper telemetry or slip ring equipment, will give an indication of the torque required to drive the pump. This approach is similar to using a torque shaft except that a longer baseplate is not required to accommodate the length of the torque shaft. However, some accuracy is sacrificed. If the pump speed is constant or known, the power required by the pump can be calculated.
It is important to know that a power monitoring device must not be used as an overpressure protection device for a positive displacement pump. The pump must be suitably protected by a pressure relief valve.
For more information on condition monitoring for rotary pumps, see ANSI/HI 9.6.9 Rotary Pumps – Guidelines for Condition Monitoring.
Some air-operated pumps are designed to be submerged. Before submerging a pump, check the chemical compatibility of the material of the wetted and nonwetted parts of the pump with the liquid in which the pump is to be submerged. Check the manufacturer’s operation manual for a particular pump before trying to use it in a submergible application.
When submerging a pump, a hose should be attached to the pump air exhaust with the exhaust piped above the liquid level to prevent the liquid from entering the pump (see Figure 10.4.1.6). To learn more about air operated pumps, see ANSI/HI 10.1-10.5 Air Operated Pumps for Nomenclature, Definitions, Application, and Operation.
A booster for a reciprocating power pump is normally a centrifugal pump but may be a positive displacement pump under special conditions (see item 4 below). Care must be exercised in the selection and installation of a booster pump, because improper selection and/or installation can result in increased pulsations and attendant problems. Here are some recommendations:1) Install a booster pump as close to the inlet source as practical;2) The booster pump must add enough pressure to the system to provide sufficient Net Positive Suction Head Available (NPSHA) to the power pump, allowing for acceleration head,friction losses and pressure pulsations due to acoustical resonances;3) Install a pulsation dampener in the inlet line adjacent to the power pump liquid cylinder. Consult dampener and pump manufacturers for proper location of the device. The dampener can often be omitted between a centrifugal booster pump and a low-speed power pump under any of the following conditions: i) Diameters of inlet and outlet connections of a booster pump are equal to, or larger than, inlet connection on power pump; ii) Diameters of all piping between liquid source and power pump are equal to, or larger than, inlet connection of power pump;
iii) The booster pump is sized for maximum instantaneous rate of flow of the power pump. The following table gives the percentage that the maximum instantaneous rate of flow exceeds the mean rate of flow for each type of power pump.
iv) Acceleration head is calculated not only between the booster and power pump but also between the liquid source and booster.4) If the booster pump is a constant-speed positive-displacement pump, a self-regulating bypass valve is required between pumps. The booster pump must be sufficiently oversized to provide the minimum flow required through the bypass valve.To learn more about reciprocating pumps, see ANSI/HI 6.1-6.5 Reciprocating Power Pumps for Nomenclature, Definitions, Application, and Operation.
The following precautions should be taken in forming orifices for pressure-measuring instruments and for making connections:
a) The opening in the pipe shall be perpendicular to the wall of the liquid passage.b) The wall of the liquid passage shall be smooth and of unvarying cross section for a distance of at least the larger of two pipe diameters or 12 in (300 mm) preceding the orifice. All tubercles and roughness shall be removed with a file or emery cloth.c) The opening shall be of a diameter d from 0.125 to 0.25 in (3 to 6 mm) and a length equal to twice the diameter.d) The edges of the opening shall be provided with a suitable radius tangential to the wall of the liquid passage and shall be free from burrs or irregularities. Figures 3.6.9a and 3.6.9b show two suggested arrangements of taps or orifices in conformance with the above.e) Manometers, when used on very viscous products with wet lines interfacing directly or indirectly to the manometer, require larger inlet taps into the pipeline to avoid excessive lag time for fluid to stabilize in the measuring system.
Where more than one tap or orifice is required at a given measuring section, separate connections, properly valved, shall be made. As an alternative, separate instruments shall be provided.Multiple orifices can be connected to an instrument, except on those metering devices such as venturi meters, etc., where proper calibrations have been made on an instrument of this form.All connections or leads from the orifice tap shall be tight. These leads shall be as short and direct as possible. For the dry-tube type of leads, suitable drain pots shall be provided and a loop shall be formed of sufficient height to keep the pumped liquid from entering the leads. For the wet-tube type of leads, vent cocks for flushing shall be provided at any high point or loop crest to ensure that there are no leaks.For additional information regarding rotary pump test procedures, see ANSI/HI 3.6 Rotary Pump Tests.
A mechanically coupled disc diaphragm liquid end (see Figure 188.8.131.52) contains a flexible, round diaphragm, clamped at the periphery, which is in direct contact with the process liquid being displaced. This type of design is inherently leak free.
The diaphragm material is typically a fluoropolymer, elastomer, or fluoropolymer-elastomer composite. A connecting rod is connected directly to the diaphragm. The diaphragm is not pressure balanced as the process pressure is acting on one side of the diaphragm and atmospheric pressure is acting on the other side. This results in higher stress levels in the diaphragm and therefore these pumps are typically used for lower pressure applications. In operation, the process liquid is admitted through the suction check valve as the diaphragm/connecting rod assembly moves away from the wet end. As the diaphragm/connecting rod assembly moves towards the wet end, the suction check valve closes and the discharge check valve opens discharging liquid.A hydraulic coupled disc diaphragm liquid end (Figure 184.108.40.206a) contains a flexible, single or double configuration diaphragm, clamped at the periphery, and is in direct contact with the process liquid being displaced. This type of liquid end design is inherently leak free. The diaphragm material is typically a fluoropolymer, elastomer, or fluoropolymer-elastomer composite. Liquid end designs featuring flexible metallic diaphragms are available and used in applications where severe operating conditions prohibit the use of fluoropolymer or other elastomers.
In operation, the diaphragm is moved by a hydraulic fluid, which is displaced by a reciprocating plunger or piston. The stresses in the diaphragm are minimal, as the process pressure acting on one side of the diaphragm is balanced by the hydraulic pressure acting on the opposite side. The process liquid is admitted through the suction check valves as the diaphragm moves rearward. As the diaphragm moves towards the wet end, the suction check valve closes and the discharge check valve opens discharging liquid. Liquid end designs of this type may include provisions such as contour plates, springs, or diaphragm positioning hydraulic control valves (Figure 220.127.116.11b) to ensure the diaphragm does not move beyond its elastic limits.For additional information regarding various controlled-volume metering pumps, see ANSI/HI 7.1-7.5 Controlled-Volume Metering Pumps for Nomenclature, Definitions, Application, and Operation.
Rotary pumps may be used to transfer slurries when the application dictates. Hard-particle slurries can cause abrasive wear to efficiency-critical components in rotary pumps, which results in reduced volumetric and mechanical efficiency. For that reason, users must carefully select and apply rotary type pumps in slurry service.
The concentration of solids usually has the greatest effect on fluid characteristics of the mixture. A fluid containing only a trace to 3 percent by weight of fine sand content in crude oil, for example, would not be considered a slurry. A range of 5 to 15 percent by weight slurry of paper stock, on the other hand, goes through three different stages of fluidity or shear characteristics, while high concentrations of finer solids can change watery carrier liquid to viscous—even thixotropic—fluid.
The size and general shape of the solids in a slurry normally determine whether a particular rotary pump configuration and/or cavity, or chamber size, can be used for the slurry in question. Typically, the cavity must be greater than particle size. In some pumps, the rotor configuration will accept larger particles as long as the size, distribution and shape are controlled. The size must be related to porting, rotor cavity size, configuration and interaction, and operating clearance.
The hardness of solids greatly influences the wear rate of the rotary pump because there is close relative motion of surfaces at the rotor tip or flank. The range of hardness is from soft solids, such as polymer chains of lubricating quality, to fibrous products that are usually soft but may carry hard particles, to very hard, irregular particles such as sand and carbon. Carbon is mainly soft, but the aggregate carries lumps with hardness approaching that of silicon carbide. A qualitative scale listing the hardness of several materials is often useful to compare solids and select pump materials (see Figure 18.104.22.168.4).Critical factors governing a system handling a watery slurry, in which the solids have a much higher specific gravity than the carrier liquid, are the settling rate and characteristics. The settling rate is a measure of how quickly solids will fall out of suspension in a slurry. Coarse solids with high settling rates that are carried in a rotary pump require many precautions to prevent bridging, draining and squeeze-out.Because speed directly affects product shear and the relative velocity of solids to the pump housing and rotor, pump speeds should always be reduced well below those for nonabrasive fluid applications, regardless of the type of rotary pump used for abrasive slurries.For more information on this subject, see the Hydraulic Institute Standard ANSI/HI 3.1-3.5 Rotary Pumps for Nomenclature, Definitions, Application and Operation.
The viscosity of a pumped fluid will affect the input power and magnetic coupling torque required in a rotary sealless pump. When selecting basic magnetic couplings for these pumps, users should consider the maximum pressure differential, speed and viscosity at which the pump will operate. These factors determine the pump’s required input power and torque. To calculate the torque requirement of the magnetic coupling, the torque required to overcome viscous friction between the inner magnet and the containment shell must be added to the pump input torque.When handling high-viscosity liquids, users must consider concerns about suction lift, port size, friction losses and reduced pump speeds. In addition to these traditional considerations, users who are selecting a sealless rotary pump should also evaluate the effect that high-viscosity liquids have on the temperature rise of the magnetic coupling. High-viscosity liquids may reduce the amount of cooling flow at a given pressure, causing diminished cooling and increased temperature rise within the coupling. Users should carefully consider the effects of high viscosity, including increased torque required and potentially elevated magnetic coupling temperature.
Starting torque for a pump that was previously primed with viscous pumpage can be a problem unless appropriate consideration has been given to coupling sizing. The use of a soft-start device may avoid decoupling on startup when pumping a high-viscosity liquid.
When pumping low-viscosity liquids, traditional concerns are poor lubricating qualities and slippage. Buyers should also carefully consider the liquid temperature, temperature rise and vapor pressure when applying a rotary sealless pump. The temperature of the liquid in the coupling should be maintained well below the vaporization point of the liquid being pumped. The viscosity of many liquids changes with shear and temperature. If viscosity increases, additional torque could be required to pump the liquid, and increased heat buildup in the drive section may occur.For more information on liquid properties, see the Hydraulic Institute (HI) Standard American National Standards Institute (ANSI)/HI 4.1-4.6 Sealless, Magnetically Driven Rotary Pumpsfor Nomenclature, Definitions, Application, Operation, and Test.
The pumping action is developed by a reciprocating piston. This reciprocating motion develops a flow profile represented by a sine wave. Actual rate of flow is determined by the following formula: Rate of flow - Displacement x Cycles per unit of time x Volumetric efficiency
Figure 7.3.1a illustrates how the rate of flow from a reciprocating, controlled-volume metering pump is minimally affected by changes in discharge pressure.
Figure 7.3.1c shows rate of flow versus stroke-length setting for a controlled-volume metering pump at a given pressure and stroking speed. The curve is linear. The curve is not necessarily proportional in that 50% stroke setting may not equal 50% flow. This is because the curve may not pass through zero on both axes simultaneously. By measuring flow at two rate-of-flow settings, plotting both points, and drawing a straight line through them, other rates of flow versus stroke length settings can be accurately predicted.
The relationship between theoretical and actual rate of flow of a controlled-volume metering pump is illustrated in Figure 7.3.1d. The difference between the theoretical and actual curves represents the volumetric efficiency (VE), the ratio of the actual pump delivery per stroke to theoretical displacement, expressed as a percent.For more information on controlled-volume metering pumps, see ANSI/HI 7.1-7.5 Controlled-Volume Metering Pumps for Nomenclature, Definitions, Application, and Operation.
Because rotary pumps are designed with close running clearances, clean piping is a must. Dirt, grit, weld bead or scale, later flushed from an unclean piping system, will damage and may seize the pump. Figure 22.214.171.124 illustrates pipe-to-pump alignment considerations.
Piping should be installed on supports independent of the pump. Supports must be capable of carrying the mass of the pipe, insulation and the pumped fluid. Supports may be hangers, which carry the mass from above, or stands, which carry the mass from below.
Clamps or brackets may be used to secure piping to existing columns. Supports must allow for free movement of the piping caused by thermal expansion or contraction. Supports should be installed at intervals that uniformly and amply support the piping load, precluding contact with adjacent piping and equipment. Pipe strains or stresses transmitted to the pump by improper piping support systems may cause distortion, wear or binding of the rotary members as well as excessive power requirements.
Piping systems that contain expansion joints must be designed so the expansion joint is not exposed to more motion than accounted for in its design. Expansion joints or flexible connectors should not be used to compensate for misaligned piping.
Threaded joints should be coated with compounds compatible with, but not soluble in, the pumped liquid. End users working with Teflon-taped joints should be careful to prevent shredded pieces of Teflon from entering the piping system. Piping should start at the pump and work toward the source of supply and the point of discharge. Shutoff valves and unions (for pumps with tapped ports) are recommended to facilitate future inspection and repair. Reducers are preferred to bushings when a change in pipe size is necessary. Avoid unnecessary restrictions in the pipeline, such as elbows, sharp bends, globe or angle valves, and restricted-type plug valves.
Users should predetermine pipe size by taking into account the required flow rate; minimum or maximum velocities; the fluid viscosity at the lowest pumping temperature; the length of the piping system, including valves, strainers and other restrictions; and the elevation of the pump with reference to supply and discharge points.
For additional information regarding rotary pump test procedures, see ANSI/HI - 3.1-3.5
Leakage from installed pumps is detected in a number of ways depending on the hazard posed by the liquid being pumped and the surrounding environment. Leakage is monitored to identify the failure of the seal or pressure boundary. These leaks may be in the form of liquid or vapor.
Pressure buildup is one form of monitoring leakage. Leakage through the inboard seal of a dual unpressurized seal arrangement may be detected by a change in pressure in the seal reservoir containing the buffer fluid. This is accomplished by blocking off the reservoir from the vent for at least 10 minutes and noting the increase in pressure. Pressure buildup in secondary containment areas of sealless pumps may also be used to indicate leakage past the primary containment.
Another form of monitoring leakage is noting a change in barrier fluid flow. Leakage through the inboard seal of a dual unpressurized seal arrangement may be detected by monitoring the gas/liquid flow from the seal to the vent system or collection system. Leakage through the inboard seal of a dual pressurized seal arrangement may be detected by measuring the change in level of barrier liquid from the circulation system and reservoir. See Table 126.96.36.199.2.5.
For a complete list of means of monitoring leakage, refer to ANSI/HI 9.6.9 Rotary Pumps – Guideline for ConditionMonitoring.
Table 188.8.131.52.2.5. Application guidelines for leakage monitoring systems’ mechanical seals
Uniform procedures for the setup and testing of controlled-volume metering pumps and for recording the test result data are available in ANSI/HI 7.6 Controlled-Volume Metering Pumps for Test.
The test procedures cover the following:
Optional testing for consideration (based on specific application or criticality of service) includes:
The items the pump test may use include, but are not limited to, the following:
Test setups that do not conform with respect to intake structure, piping and measuring equipment may not duplicate test facility results.
For additional information on controlled-volume metering pump testing, refer to ANSI/HI 7.6 Controlled-Volume Metering Pumps for Test.
The impact of the fluid type on a rotary pumps is generally viscous related. The viscosity of a fluid is the tendency of the fluid to resist an internal shearing force. The viscosity of the fluid being pumped affects the net positive inlet pressure required (NPIPR) and will impact the pump flow slip.
A fluid is Newtonian when the ratio of shear stress to shear rate is a constant for all shear rates, is independent of time, and zero shear rate exists only at zero shear stress (see Figure 1). Most mineral oils at temperatures above the cloud point (the temperature at which the oil begins to appear cloudy), solvents and water approximate this condition and are considered Newtonian fluids. The viscosity of these fluids is independent of rate of shear.
A non-Newtonian fluid will change viscosity with changes in the rate of shear applied to the fluid and/or the length of time at shear.
Several types of non-Newtonian fluids are defined in Figures 2 through 6. When the ratio of shear stress to shear rate increases as shear rate increases, reversibly and independent of time, a fluid is said to be dilatant (see Figure 2). Highly concentrated pigment-vehicle suspensions—such as paints, printing inks and some starches—are dilatant fluids. The apparent viscosity of these fluids increases as the rate of shear increases. Some dilatant fluids solidify at high rates of shear.
When the ratio of shear stress to shear rate is constant for shear rates above zero, it is independent of time, but when shear occurs only for shear stress above a fixed minimum greater than zero, a fluid is termed plastic. A plastic fluid, such as putty or molding clay, is characterized by a yield point. This means that a definite minimum stress or force must be applied to the fluid before any flow takes place.
When the ratio of shear stress to shear rate decreases as shear rate increases, reversibly and independent of time, and zero shear rate occurs only at zero shear stress, a fluid is pseudo-plastic. Many emulsions, such as water-base fluids and resinous materials, are pseudo-plastic fluids. Their apparent viscosity decreases with increasing shear rates but tends to stabilize at high rates of shear.
A fluid is thixotropic when the ratio of shear stress to shear rate decreases and is time-dependent in that this ratio increases back to its “rest” value gradually with lapse of time at zero shear rate and stress, and decreases to a limit value gradually with lapse of time at constant shear rate. Most greases, drilling mud, gels and quicksand are thixotropic fluids when the apparent viscosity of these materials decreases for an increasing rate of shear and for an increasing length of time at shear.
When the ratio of shear stress to shear rate is constant for all shear rates at any given instant of time, but increases with time, a fluid is rheopectic. Some greases are intentionally manufactured to have partial rheopectic properties that facilitate pumping in a stable condition; however, upon shearing in a bearing, the grease builds up to a higher apparent viscosity.
Additional information about fluids and the effect of viscosity on pump and system performance may be found in ANSI/HI 3.1-3.5 Rotary Pumps for Nomenclature, Definitions, Application, and Operation.
Because of the characteristic pulsating flow of metering pumps (see Figure 184.108.40.206), where peak flow rates can reach three times the average flow, operators must carefully consider suction piping to ensure that it can deliver adequate fluid to the pump inlet
When connecting more than one pump head to a single suction manifold, operators must consider a number of issues. Suction manifold piping must accommodate peak demand of the pumps throughout its full range of operation, as well as prevent offgassing of liquids with high vapor pressure or dissolved gases. This is accomplished by appropriately increasing the diameter of the suction manifold and pump head connections or by adding accessories to increase flow of fluid to the pump inlet. The most important consideration is whether the multiple heads are part of a multiplex pump (multiple heads connected to a single motor or driver) or if the heads all have independent motors or drivers. In most cases, multiplex pumps are connected to a common manifold piping arrangement, driving these recommendations. The other option is to connect each pump head to its own suction line connected to the supply source.
In a multiplex where a single driver is used, if the pumps all operate from the same gear set or with multiple gear sets operating at the same stroking speed, industry best practices suggest that the drive mechanism for each individual pump head should be run out of phase (i.e. duplex pumps, 180 degrees; triplex pumps, 120 degrees; and so on). This staggers the peak suction requirements of the individual pumps. Running several pumps from the same driver (motor) with proper phasing of the drive elements to ensure this timing during all running conditions and the best suction flow characteristics as they relate to piping design.
When independent pumps, each with its own driver, are connected to a common suction line, or if the application requires multiplex pump heads operating in phase, operators must make allowance in the piping design in case all pumps simultaneously demand peak suction flow.
For more information on piping guidelines for control-volume metering pumps, refer to the new standard ANSI/HI 7.8 Control-Volume Metering Pump Piping Guideline.
Numerous applications in a combined-cycle power plant require the delivery of precisely metered quantities of chemicals. Chemical metering pumps are positive-displacement types (see Figure 6.16) and are designed for continuous operation.
One of the key applications is to deliver chemical solutions to control the amount of emissions. Natural-gas fired, combined-cycle power plants produce approximately one half of the CO2 emissions and one third of the nitrogen oxides (NOx) produced by conventional power plants. They also eliminate sulfur dioxide emissions.The combustion turbines in a combined-cycle plant produce low NOx and CO2 emissions. The heat recovery steam generator then uses a selective catalytic reduction scrubber system to reduce the NOx to a very low level (approximately 30 percent of a simple-cycle combustion turbine's NOx emission using similar technology). Some are as low as 10 percent of the simple-cycle emissions.The scrubber system is engineered to treat a specific cubic feet per minute range of gas, within a specific temperature range. Further, the system is engineered to reduce the level of contaminants from a specific inlet concentration to a specific outlet concentration. Other typical metering pump applications include boiler feed, cooling tower feed, wastewater treatment, condensate treatment and feeding the makeup water system.
More information about the use of chemical metering pumps in combined-cycle power plants can be found in Power Plant Pumps: Guidelines for Application and Operation to Maximize Uptime, Availability, and Reliability.
No fixed rules are available for selecting the pump type for a boiler feed application. Pump selection depends on a combination of the manufacturer’s application experience and product offerings and the end user’s experience in maintainability and reliability. If the plant will be cycled instead of base loaded, then special consideration must be given to the use of a sufficiently robust and proven pump design.Electric utilities favor the between-bearings pump arrangement for the highest level of reliability. For pressures up to approximately 1,800 psi (125 bar), typical boiler feed pumps include horizontal multistage ring-section (type BB4, see Figure 6.1) or axial split-case designs (type BB3), although in some applications there may be a preference for a double-case pump (type BB5).
For higher pressures, the boiler feed pump types used are usually multistage ring-section (type BB4) or double-case pumps (type BB5).Some end users choose to give particular consideration to the main pump construction features. They include:Type BB4 ring-section and type BB5 double-case pump types both have radial sealing joints, which are geometrically simpler than an axial split-case pump sealing joint. The double-case pump has fewer sealing joints than the ring-section pump.Type BB3 axial split-case pumps and the axial split inner casing version of the type BB5 double case pump allow for the entire assembled rotor to be removed and balanced as a unit. These rotors usually have opposed impeller arrangements that reduce the hydraulic thrust carried by the thrust bearing.The ring-section and double-case pump geometries inherently lend themselves to design for pressure containment. The double-case pump type has an integrally safer containment for pressures exceeding 3,600 psi (250 bar).
For more information about boiler feed pumps and their application in combined-cycle power plants see HI’s newest guidebook Power Plant Pumps: Guidelines for Application and Operation to Maximize Uptime, Availability, and Reliability
Boiler circulating pumps circulate water within the boiler to enhance boiler operation. They take suction from a header connected to several downcomers from the bottom of the boiler drum and discharge through additional tube circuits. This means the water pumped is at boiler temperature and pressure.Boiler circulating pumps are designed for high temperature (usually between 150 and 315 °C [300 and 600 °F], depending on boiler size and rating), and high pressure (corresponding to boiler temperature and water vapor pressure). For small boilers, with relatively low temperatures and pressures, conventional overhung pump designs (such as HI pump type OH2 seen in Figure 3.11) may be suitable for boiler circulating service.
Boiler circulating pumps must develop only enough head to overcome the friction of the tube circuits. However, the combination of high temperature and pressure results in sealing conditions that require special sealing devices.Because of the relatively low head requirements, the pumps are single stage with single suction impellers and a single seal chamber. This creates a problem of unbalanced axial thrust, which may require special pump bearing systems or balancing arrangements.An alternate solution is to use pumps of wet motor construction, where the pump and the motor are inside the pressure vessel, eliminating sealing and unbalanced axial thrust issues. Such special pumps are welded into the boiler piping. For higher temperatures to 365 °C (685 °F) and pressures from 124 to 193 bar (1800 to 2800 psi), special pump designs are required.For more information about boiler circulating pumps and their application in combined-cycle power plants see HI’s newest guidebook Power Plant Pumps: Guidelines for Application and Operation to Maximize Uptime, Availability, and Reliability.
Auxiliary cooling water pumps (see Figure 6.15) usually operate in a closed loop system using high-quality water. The water received from the circulating water pumps circulates through the tube side of the auxiliary cooling water heat exchanger. The auxiliary cooling water circulates through the shell side and is cooled. The auxiliary cooling water then proceeds to various small heat exchangers used for the cooling of pump and motor bearing frames, mechanical seals, etc.
The fluid pumped is clean, high-quality water, as opposed to the main cooling system, which takes water directly from the cooling tower or other source. The reason for using clean, high-quality water to cool pump, fan and motor bearing frames, mechanical seals, heat exchangers, etc. is that they could be easily plugged if the cooling water contains high levels of suspended or dissolved solids. The fluid temperature usually ranges from 10 to 27 °C (50 to 80 °F).Auxiliary cooling water pumps may be either single or multistage, depending on the pressure and flow requirements of the particular system. Single stage end suction pumps of the heavy-duty process types OH1, OH2, and BB1 are popular for both low- and medium-pressure service. For higher pressure, multistage pumps VS6 types are used. The castings (metallurgy) for subject pumps can be all cast iron or ductile iron construction. However, for VS6, the castings would be carbon steel. A typical upgrade for these pumps would be carbon-steel castings with 12% chrome stainless impeller and wear rings.For more information about the pump types used in combined-cycle power plants, see Power Plant Pumps: Guidelines for Application and Operation to Maximize Uptime, Availability, and Reliability.
The operating range of the plant compared to the operating range of the pump, as well as reliability considerations and first cost, will determine how many pumps are required. Large thermal power stations typically have two or three installed 50% boiler feedwater pumps.The recommended margin of Net Positive Suction Head Available (NPSHA) over Net Positive Suction Head Required (NPSH3) for boiler feed pumps can be relatively high. This will depend on rate of flow, the operating speed of the pump, whether the first-stage impeller is a single or a double suction type, and the expected operating range of the pump. The pump manufacturer will provide NPSH3 and margin recommendations to mitigate cavitation and provide long life of the first-stage impeller.A key factor in pump selection for demanding boiler feed service is rotor hydraulic axial balance. Two types of rotor construction need to be considered: in-line (tandem) impellers (see Figures 6.2 - 6.4) and opposed (back-to-back) impellers (see Figure 6.6). With in-line impellers, the hydraulic axial thrust of the impellers is usually countered by a balancing device, either a balance disk or a balance drum, or sometimes a combination of both. The advantages of a balance disk design may include reduced internal recirculation (better volumetric efficiency) and the ability of the disk to compensate for wear and the amount of thrust that requires balancing. The disk is referred to as a self-compensating balancing device, and may be used to exclude an axial thrust bearing in certain designs, thereby reducing initial cost and simplifying rotor construction and assembly. The balance drum does not compensate for changes in thrust and typically exhibits more internal recirculation losses than a balance disk.
The advantage of the balance drum over the balance disk is in the openness of its running clearance. For pumps in systems where severe upsets may occur, or where foreign material is continuously present, balance drums (with their larger running clearances) may prove less sensitive (less susceptible to damage) than balance disks.For more information about boiler feed pumps for combined-cycle power plant use, see HI’s newest guidebook Power Plant Pumps: Guidelines for Application and Operation to Maximize Uptime, Availability, and Reliability.
Many pumping system components are not dynamic. That is, these components allow fluid or heat transfer but, aside from thermal expansion or structural vibration, do not move and do not have dynamic surfaces that wear out. The most common types of problems in these components are leakage, fouling, valve failure and cracks in pipe supports.For many industrial facilities, the energy consumed in pumping fluids comprises a large fraction of the facility’s overall electricity usage. Operators of pumping systems often are not aware of how effectively the energy required to pump the fluids is being used. If they are unable to meet their functional requirements, a large part of the plant, even the entire facility, may be affected.A few symptoms in pumping systems often indicate if improper sizing, selection, operation or other issues affect the system’s performance. They are:
The existence of any one of these symptoms does not guarantee significant savings potential, but it indicates a greater likelihood of achieving savings. The more symptoms present, the greater the likelihood of potential energy savings.
See the Hydraulic Institute’s guidebook Optimizing Pumping Systems: A Guide for Improved Energy Efficiency, Reliability & Profitability for more information.
Dual seals include more than one set of sealing surfaces. Typically dual seals require the use of an external liquid or gas and are operated with an auxiliary system. As compared to a single seal, they require more auxiliary systems and monitoring. Their total initial cost is usually higher and require the installer to have additional training. The benefits are increased reliability and safety. Dual seals are subcategorized as dual pressurized and dual unpressurized.In a dual pressurized seal, the external fluid is at a higher pressure than the seal chamber pressure and thus lubricates all seal faces of the mechanical seal. In this design, the external liquid is called a barrier fluid. In the dual unpressurized seal, the external fluid is at a lower pressure or, as in most applications, at ambient pressure. The external fluid is called a buffer fluid. Dual seals come in several configurations and with a variety of application-specific design features. Their faces can be arranged face-to-back, face-to-face, and back-to-back.Dual seals are used when one or more of the following criteria cannot be achieved by using a single seal design:
The advantages of deal seals include:
The advantages of the pressurized dual seal over unpressurized version are:
For more information on mechanical seals, see the HI Guidebook Mechanical Seals for Pumps: Application Guidelines.
Harmonics are undesirable voltages and currents that may appear on electrical power systems as a result of certain types of nonlinear loads, such as computers, electronic lighting, welding supplies, uninterruptible power supplies (UPS) and variable frequency drives (VFDs). When installing VFDs, operators should consider the existing harmonic content and account for any harmonic impact the newly installed VFD may have on the electrical system or the electrical utility. Harmonics can be a concern for any size VFD, whether it is a 500-horsepower (HP) unit in a building or a 75-HP unit in an irrigation application. Operators should review the existing installation—not just the current project—with their consultants and drive representatives in order to find the most positive installation solution.A full-wave bridge rectifier or converter converts alternating current (AC) power to direct current (DC) power, no longer a sinusoid waveform, and the capacitor bank acts as a filter smoothing this converted voltage. All AC/DC converters used in different types of electronic systems can increase harmonic disturbances by emitting harmonic currents directly into the electrical grid.Harmonic distortion can cause overheating of supply cables and distribution transformers (reduced insulation life), LED and lighting disturbance, nuisance tripping of circuit breakers, ancillary electronic equipment damage and potential faulty readings, and damage of backup generators.An active filter cancels out most harmonics on the current line. The active filter injects equal and opposite harmonic current onto the line in a similar way that noise-canceling headphones inject noise-canceling signals into the air around the wearer’s ears. As the noise-canceling signal from the headphone results in the cancellation of noise, the active filter cancels out the harmonics, as noted in Figure 2.12.
For more information on variable frequency drives, see the Hydraulic Institute Guidebook Variable Frequency Drives: Guidelines for Application, Installation and Troubleshooting.
For pumping equipment to perform a long-term task, the end user must have a design suitable for the intended task. The equipment assembly also must be operated and maintained within the design and operating duty condition limits. The majority of the reliability shortfalls are due to operation outside the design limits or abnormal operating conditions.
All pumping equipment components have a design life for a given set of operating conditions. The pumping equipment design chosen for a particular service and application can determine the design life and, to varying degrees, the product reliability.
The life of a piece of pumping equipment is shown in Figure 3.2. If the actual product is defect-free, to specification and the specified operating conditions are kept, design life for the product will be met. However, if the materials or design are deficient, the operating conditions are different or more difficult, or the product is operated or maintained improperly, reliability will be impaired and operating life will be reduced.
For more information, check out Hydraulic Institute’s newest guidebook Reliability of Pumping Equipment: Guidelines for Maximizing Uptime, Availability, and Reliability.
The type of pump that can be used depends on the location of installation. These location considerations include the size of the installation space, discharge piping alignments, wet-well depth, interior/exterior installation and power.
The area where the pump is installed must allow space for the pump and driver assembly to be properly mounted, for the inlet and discharge piping and valves, and for the necessary power wiring and switches. Operators and maintenance personnel must have enough space to access the pump and complete any required work or removal of the pump. Because their installation space is significantly different than that of other pumps, horizontal and vertical pump assemblies require special evaluation.
New designs allow designers to create the space required to meet the new pump requirements. Retrofit designs may not afford this ability. The existing structure, wet well, hydraulics, control sequencing or installation requirements may become the primary considerations when end users are choosing a pump. The fluid being pumped can create specific requirements for the driver and/or the power distribution components that serve the pump assembly. These requirements can change the sizing of the components, so users should also evaluate these factors when installing the pump assembly.
For more information on wastewater treatment pumps, see HI’s newest guidebook Wastewater Treatment Plant Pumps: Guideline for Selection, Application, and Operation.
The primary treatment step physically separates insoluble solids from floatable materials contained in the wastewater. For example, primary sludge consists of the solids generated from the primary sedimentation process. The primary sludge may be transferred to a solids-processing system, such as a gravity sludge thickener, where the solids are concentrated and pumped to the solids-handling facilities. The primary sludge stage can incorporate many types of pumps, including horizontal and vertical solids-handling rotodynamic, submersible solids-handling rotodynamic, piston, progressive cavity and rotary lobe.
When incorporating a solids-handling rotodynamic pump (see Figure 4.3), end users must consider the change of sludge consistency because the most concentrated sludge is pumped in the beginning of a pumping cycle. When most of the sludge has been pumped, diluted sludge with consistency and hydraulic characteristics similar to water will be pumped. To determine the required motor power for a rotodynamic pump, users should plot system curves and pump curves for both the most dense sludge and water.
A progressive cavity (PC) pump (see Figure 4.18) can be used in the primary treatment stage and will deliver constant flow regardless of pressure. The flow is controlled by the pump speed. For low-solids contents, full speed is typically allowed. If sand and grit are in the sludge, the speed should be adjusted down to reduce wear. Because of their design with a metallic rotor in direct contact with a stationary elastomeric stator, PC pumps must not run dry. If using a PC pump, install a pressure guard that shuts the pump off when it experiences a pressure drop indicating dry run.
VFD units are used to vary motor speed, which allows the pump to provide different rates of flow to meet system requirements. They also provide an option to reduce inrush current. They have time-adjustable ramp-up and ramp-down features to control motor acceleration and deceleration times. The controller may be programmed to electrically disconnect the motor on signal shut-down. VFDs can provide numerous functions in wastewater pump applications, which include:• Regulate the operation of pumps or the pump station in relation to wet-well depth, discharge flow or force main pressure, which can potentially eliminate the need for control valves• Provide energy-efficient operation of pumps by matching best efficiency operating points and/or matching pump operation to varying demand loads• Ability to over-speed pumps to meet certain specific short-term application conditions• Provide power phase conversion (producing three-phase power from a single-phase power supply)• Create power at a different frequency (e.g., producing 60-Hertz [Hz] power when only 50 Hz is available)• Adjust the pump operational speed (or torque) instead of using a control valve to restrict pump output, which reduces downstream flow or pressure• Reduce mechanical stress of conditions such as cavitation and water hammer on both the pump and pump system
To minimize the risk of clogging the pump impeller, users should consider alternating pumps to create back-flush and to bring them to full speed before stopping or at the start of a pumping cycle. Operating pumps at low speeds or with slow increases and decreases in speed can cause pump clogging.
Boiler circulating pumps circulate water within the boiler to enhance boiler operation. They take suction from a header connected to the bottom of the boiler drum and discharge through additional tube circuits. This means the water pumped is at boiler temperature and pressure. For this reason, standard boiler circulating pumps must be designed for high temperature (usually between 300 and 600 degrees F [150 and 315 C] depending on boiler size and rating) and high pressure (corresponding to boiler temperature and water vapor pressure). For small boilers, with relatively low temperatures and pressures, conventional overhung pump designs may be suitable for boiler circulating service.
Boiler circulating pumps must develop enough head to overcome the friction of the tube circuits; however, the combination of high temperature and pressure results in sealing conditions that require special sealing systems. Because of the relatively low head requirements, the pumps are single-stage with single-suction impellers and a single seal chamber. This creates a problem of unbalanced axial thrust, which may require special pump bearing systems or balancing arrangements. An alternate solution is to use pumps of wet-motor construction, where the pump and the motor are inside the pressure vessel, eliminating sealing and unbalanced axial thrust issues. Such special pumps are welded into the boiler piping. For higher temperatures to 685 F (365 C) and pressures from 1,800 to 2,800 pounds per square inch (psi) (124 to 193 bar), special pump designs are required.
For more information about pumps used in combined-cycle power plants, see HI’s guidebook Power Plant Pumps: Guidelines for Application and Operation to Maximize Uptime, Availability, and Reliability.
The optimal performance or life of pumping equipment will be achieved when the equipment is correctly designed, sized and selected for a given application and the units are operated (and maintained) at the specified conditions.
The design life can be different for varying services and equipment types. In any application, different shareholders in the overall outcome will influence the type of equipment selected and the way it is applied and maintained. Main shareholders and influencers in the design, selection, application and maintenance of the equipment include the following:
In order to obtain maximum reliability and mean time between repairs (MTBR), pump equipment must be properly designed and selected for the intended application, and it must be operated according to the manufacturer's recommendations. Actual equipment life may not match the predicted value, which depends on the different ways of measuring and analyzing reliability.
Comon pump types for wastewater processes are vertical sump pumps or submersible pumps. In a controlled process, the estimated or mean time between repairs is typically two to eight years, depending on the severity of service for both types of pumps. For pumps that require pressures up to 285 pounds per square inch gauge (psig) and temperatures up to 450 F and are exposed to a harsh abrasive environment, the estimated or mean time between repairs is typically two to four years.
If in practice the equipment life falls short of the industry averages, then a thorough review should be conducted.
For more information on the reliability of pump applications, reference HI’s guidebook Reliability of Pumping Equipment: Guidelines for Maximizing Uptime, Availability and Reliability.
Some important considerations are altitude, temperature of the liquid and speed of operations.
For more information on water and wastewater pump selection recommendations, refer to HI’s Guidebook Wastewater Treatment Plant Pumps: Guidelines for Selection, Application, and Operation.
Sealing devices can be classified as static seals, such as gaskets, and dynamic seals, such as rotating interfacial axial seals, also known as end face mechanical seals. Mechanical seals are used in many types of pumps of various sizes and pressure ratings and are used to transport many different fluids in numerous industries. Therefore, mechanical seals are available in a variety of configurations. Their selection depends on the application conditions; however, regardless of the service conditions, all mechanical face seals operate on the same basic principle. A simple mechanical seal is shown in Figure 3.1.
A seal consists of two sealing rings, either of which rotates relative to the other. One of the rings is mounted rigidly and the other is arranged so that it can move freely and align radially, axially and angularly with the rigidly mounted ring. A dynamic seal is achieved when the two rings contact perpendicularly to the pump shaft. These rings are called seal faces, one the rotating face and the other the stationary face.
The faces are lapped flat, which results in very low levels of leakage while providing long life on the basis of normal wear. Besides the two faces, the mechanical seal contains a set of secondary sealing elements and several metal parts. These serve the function of sealing dynamically and statically, loading the faces and transmitting rotation. The secondary seals provide sealing between the seal faces and the metal parts, such as mating ring housing, sleeve and gland. The metal parts transmit the torque and provide an axial mechanical force by means of a spring element to load the lapped faces.
For more information on mechanical seals, refer to HI’s guidebook Mechanical Seals for Pumps: Application Guidelines.
feed systems supply chemicals for various stages in a water or wastewater
process. The treatment process requires optimal synthesis of new biological
growth so the addition of chemicals may be required to maintain a favorable
It is also important to maintain the pH of the wastewater in the process
and sufficient alkalinity to provide for this operating pH range. As a
result, it may be necessary to add either acid or caustic for pH control,
ferric chloride or alum for nutrient removal, or supplemental carbon sources
such as methanol, glycine or acetic acid for process development.
Each chemical feed system is different depending on the type of chemical to
be pumped, its concentration, and the necessary feed rate. Small feed rates
will require metering-type pumps that can provide a specific dose of the
chemical to the receiving stream (see Figure 4.14). Metering pump flow rate
can be manually adjusted by a micrometer dial, varying the speed of the
drive, or a combination of the two. Flow rate can also be controlled
automatically in response to a process signal.
Users should be careful when
selecting a metering pump for this stage of the wastewater treatment process
because many disinfectants, such as chorine, are degassing chemicals. Air
bubbles produced by a degassing chemical can travel through the suction line
into the pump’s head. Once the bubbles are trapped inside the head, the pump
may lose prime or gas lock because of the bubbles. Metering pumps with
auto-degassing capabilities are preferred to prevent gas-lock situations. A
gas-lock situation could prevent accurate and reliable injection of the
disinfectant parameter. Corrosion inhibitors may also be used to prevent
damage on the piping and equipment.
Except with chemicals such as quick
lime, abrasion is minimal because there is no grit in most of the chemicals.
However, feed concentration rates and chemicals must be reviewed to ensure
the pump and piping system materials do not corrode when the full
concentration or diluted concentration of the chemical is applied.
Selecting the appropriate metering
pump for a wastewater application usually requires an examination of the
required chemical injection rate, which is normally determined based on
wastewater analysis. Once the injection rate is known, determine the maximum
and minimum rate of flow (cubic meters per second or gallons per hour),
pressure rating (kilopascals or pounds per square inch), desired control
options (manual, timer, pulse, analog or on-off), materials of construction
(stainless steel, PVC, PVDF, PP, FKM, EPDM, ceramic, etc.), and type of
connection (half-inch NPT, quarter-inch tubing, etc.).
For additional information on controlled-volume metering pumps in
wastewater treatment, refer to Hydraulic Institute Wastewater Treatment Plant
Pumps: Guidelines for Selection, Application and Operation.
primary purpose for using a VFD in a pumping application is to vary the speed
of the pump to optimize system efficiency and/or process control. Adjusting
the frequency applied to the motor allows the speed of the motor and pump to
be changed, usually to a lower speed than the base revolutions per minute
(rpm) stated on the motor nameplate or lower than the speed the pump would
operate with utility power.
Operating a pump at reduced speeds is usually acceptable within a specified
range, but operators must consider certain mechanical design features such as
bearing lubrication and natural frequency interactions. Anyone considering
reduced speed operation should consult the pump original equipment
Operating the motor at lower speeds usually results in a reduction of the
motor cooling system capability, because motor cooling fans produce less
airflow over the motor at reduced speeds. As a result, operators must ensure
that the motor can withstand the added heat that is created by operating
below base speed. In addition, the operating speed range should be defined
and applied to motor capability requirements. In many cases, the pump and
motor can be sped up to 10 percent above base speed without adverse effect.
Before ramping up speed, however, operators should consult the OEM to
evaluate electrical and mechanical design considerations.
In rotodynamic pump applications,
the load is governed by the pump affinity laws, which state that a change in
torque is proportional to the square of the speed change and that change in
horsepower is proportional to the cube of the speed change. This type of load
is known as a variable torque load (VT). Most motors are capable of operating
10 times slower than the nameplate rpm of the motor (10-to-1 speed range or
turndown) when driving a rotodynamic pump.
Positive displacement pumps present
higher relative loads at reduced speeds, so reduced operating speeds will
have a more significant impact on the motor and drive. Most motors can
operate with a 2-to-1 turndown speed range with this type of load, which is
commonly considered a constant torque (CT) load.
For best results, define the load
type and operating speed range in advance. Provide this information to the
pump, motor and drive manufacturers and discuss it with them to ensure the
motor and drive selection is the best fit for the pumping application.
For more information on VFDs, refer
to HI’s guidebook Variable Frequency Drives: Guidelines for Application,
Installation, and Troubleshooting.
wastewater applications, the pump driver can be a motor- or engine-driven
unit. The most common pump driver in a wastewater treatment plant is an
electric-powered induction motor (see Figure 5.1). Engine-driven pumps are
typically used in emergency or standby service when electrical power supply
is not available. Synchronous motors can also be used, but they are normally
provided for kilowatt requirements above 750 (horsepower requirements above
The selection of a type of motor,
the method used to start the motor and considerations for operating the motor
are based on many factors. These include location, atmospheric designation of
the location, operational conditions, size of the motor to be used, number of
motor starts required and power availability.
Electric motors have optional enclosure types. When choosing the type of
enclosure, operators should consider the application and the designation of
the area of use. The options for motor enclosures include open drip-proof
(ODP), totally enclosed fan cooled (TEFC) and totally enclosed nonventilated
ODP motors are suitable for general
industrial use in outdoor or indoor locations with relatively clean
atmospheres and nonexcessive ambient temperatures. The ventilation openings
of an ODP motor are constructed to prevent interference with motor operation
when liquid or solid particles less than 50 millimeters (mm) (1.97 inches)
strike or enter the enclosure at an angle up to 15 degrees downward from
vertical. In an open motor, air circulates freely through the motor for
cooling. Openings are limited in size to prevent contact with hazardous
internal parts. This provides some protection from drip, splash and falling
objects. The motor is susceptible to moisture and contamination.
A TEFC motor is enclosed to prevent
the free exchange of air between the inside and outside of the case but not
sufficiently enclosed to be airtight. The exterior surface is cooled by an
external guarded fan attached to the motor shaft, which circulates air over
the frame housing. Motor cooling is dependent on motor speed. A finned
cast-iron framed motor or a steel motor on a smaller motor frame are usually
suitable for this type of enclosure. Because air does not pass freely through
the motor, the TEFC enclosure protects the motor from harsh environments.
These motors are typically well-suited for harsh, dirty and wet environments,
and they can be operated indoors or outdoors.
A TENV motor is a frame
surface-cooled unit that is cooled by free convection. It is enclosed enough
to prevent the free exchange of air between the inside and outside of the
case, but it is not completely enclosed to be airtight. It is not equipped
for cooling outside the enclosed parts. No external fan is used; the motor is
cooled by free convection and radiation only. A TENV motor is suitable for
use in dusty, dirty locations and in humid environments. In general, it is
limited to low horsepower ratings or limited time duty rated motors.
For a full list of motor enclosures
and more information on pump drivers in a wastewater application, refer to
HI’s guidebook Wastewater Treatment Plant Pumps: Guidelines for Selection,
Application, and Operation.