Pump FAQ’s by Hydraulic Institute (UK)

Pump FAQ’s by Hydraulic Institute (UK)

FAQ’S OF PUMPS

Pump FAQsÂ® September 2001 Article #3

Q.

We have many centrifugal pumps, which take their inlet from an overhead tank. When the centrifugal pump draws down the liquid level in the tank, a vortex forms in the liquid and air is drawn into the pump. How can this problem be corrected?

A.

At low liquid levels, the formation of a vortex is dependent on the liquid velocity entering the pipe. Increasing the effective pipe area can reduce this liquid velocity. Installing a baffle plate above the pipe opening will simulate a larger intake area thereby reducing the vortex formation. See the figure below:


Q. Our plant is operating a system with several centrifugal pumps in parallel that exhibits waterhammer problems when one of the pumps is shut down. The discharge pipe from the pump is 18 inches diameter, forty feet long up to the common header. The backflow after shutdown causes severe banging of the check valve. What can be done to relieve this problem?

A. Much has been written on this subject, including a section in the â€œPump Handbookâ€ published by McGraw Hill which include fifteen pages on �waterhammer plus twelve references. Some solutions that have been applied include:

Close to the pump isolation valve before shutting down the pump. Add a flywheel to the pump in order to slow its rate of coast down. Add a surge tank or air chamber which will absorb discharge pressure

energy. Use a fast closing check valve which will close before reverse flow begins.

Use a slow closing check valve which will allow the reverse flow to shut down more slowly and avoid the shock from the sudden stop of the backflow.


Q. Many manufacturers do not extend the NPSHR curve on their pump rating curve below about thirty percent of its best efficiency point. Can the NPSHR curve be extrapolated back to shut off in order to get this information?

A. Extrapolation of the NPSHR curve back to shut off is dangerous, since different pumps act differently at very low flow rates. Some pumps exhibit a dramatic increase in NPSHR at very low flows. In addition, the condition

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known as suction recirculation becomes very pronounced at very low flows causing vortices in the impeller and localized pitting damage to the impeller blades.

The subject of low flow operation is discussed in ANSI/HI 9.6.3-1997 Centrifugal and Vertical Pumps - Allowable Operating Region.


Q. In our process, the primary circulating pump occasionally becomes airbound due to entrained air in the system and stops pumping. How much air can a centrifugal pump handle, and what can be done to prevent airbinding?

A. Centrifugal pumps can usually handle up to about five percent by volume of air. Above that, they will easily become airbound, especially at flow rates below the best efficiency point. Even when pumping five-percent air, the pressure developed will be reduced due to the reduced specific gravity of the fluid mixture. In order to avoid airbinding, try the following:

Minimize air entrainment by any means possible. Add a bypass pipe to maintain a high rate of flow.

Install a vent in the top of the suction pipe to vent air back to the source.


Q. We are operating a multistage centrifugal pump in our process that is producing too much head. We intend to reduce the pump head by reducing the impeller diameters. Can we reduce head by trimming only one or two impellers, and is it better to cut the impeller shrouds as well as the impeller vanes?

A. It is not necessary to keep all of the impellers at the same diameter. However, it is better to limit the extent of impeller reduction to five percent or less due to loss in efficiency as the impeller is cut down. You should evaluate the possible loss in efficiency by cutting only one or two stages if the reduction is more than five percent.

Regarding cutting of impeller shrouds, the advantage is that the full shroud diameter will minimize pump instability when it is operated below fifty percent of its best efficiency point. However, cutting the shrouds will reduce the impeller disc friction and improve efficiency.

Pump FAQsÂ® October 2006 Article #2

Q. I keep hearing about suction recirculation but cannot find a good explanation.

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Can you help me?

A.When centrifugal pumps are operated at rates of flow below their best efficiency point (BEP), the excess flow produced by the impeller is recirculated on the suction side of the impeller as shown in figure 5-12. These eddy currents cause local vortices on the impeller vanes, which in turn cause cavitation resulting in noise, vibration, and damage to the impeller vanes

This action usually begins between 70% and 50% of the BEP flow and many pump manufacturers do not allow operation below the onset of recirculation. Furthermore, some impellers, which are designed for low NPSH Available applications, may begin the recirculation mode close to the BEP. An example of this might be condensate drain pumps in electric power plants.

Pump FAQsÂ® November 2004 Article #1(1)

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Q. The first answer in the September issue of Pumps and Systems says that spring mounted baseplates can also work well to avoid pipe strains on pumps. How can I know if the baseplate stiffness is sufficient to work without grouting?

A. Whether to grout or not depends on the specific application and the design of the baseplate. Applications that undergo wide temperature variation may do well with floating baseplates in order to allow for movement of the pump and thereby minimize pipe strain caused by thermal expansion of the pipe. However, the baseplate in such cases must be sufficiently rigid and designed to avoid misalignment between the pump and driver shaft as the baseplate moves.

The purpose of a baseplate is to provide a foundation under a pump and its driver that maintains shaft alignment between the two. This baseplate must allow for initial mounting and alignment of equipment, survive handling during transportation to the installation site, be capable of being installed properly with minimum difficulty, allow final alignment of the mounted equipment, and allow removal and reinstallation of equipment. It must be recognized that it is not necessary that an absolutely rigid baseplate be designed to meet these requirements. At the same time, the baseplate must not be permanently deformed after the equipment is mounted at the manufacturing facility. Compliance with these design criteria, in conjunction with proper installation procedure, will contribute significantly to meeting the functional requirements.

Baseplates may be designed to be installed free standing, or to be installed using grout. A free standing baseplate must be rigid enough to maintain coupling alignment when subjected to loads from piping or motor torque. The rigidity shall prevent no more than 0.25 mm (.010 inch) parallel coupling misalignment and .005 mm/mm (.005 inches/inch) angular misalignment when subjected to maximum motor and piping loads simultaneously.

A grouted baseplate relies on the grout for the majority of its stiffness but it must be sufficiently rigid to permit handling and allow installation. It is recommended that this type of baseplate have torsional stiffness of 1.13x10 4

Nâ€¢M/rad (105 inâ€¢lb/rad).

Additional information and calculation of baseplate stiffness can be found in Appendix A of ANSI/HI 1.3-2000 Centrifugal Pumps for Design and Application.

Pump FAQsÂ® May 2006 Article #3

Q. How does a change in liquid viscosity effect the performance of a rotary

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pump? I have seen a viscosity correction graph for the performance of centrifugal pumps but nothing for rotary pumps.

A. Because of the variety of rotary pump designs, no general viscosity correction graph has been developed. Most pump manufacturers produce a range of performance curves for each pump at different levels of viscosity.

There are some general guidelines that can be used as follows:

Net positive inlet pressure required (NPIPR) increases with increasing viscosity.

Required pump input power increases with increasing viscosity. Maximum allowable pump speed decreases with increasing viscosity. Pump internal leakage or slip decreases with increasing viscosity.

Pump volumetric efficiency increases with viscosity, but pump efficiency peaks at about one thousand SSU viscosity due to the increasing viscous drag at increasing viscosity. See figure 3.31 below.

For more information see the HI standard ANSI/HI 3.1-3.5 Rotary Pumps


Q. What is reverse runaway speed? What is the cause? Is it dangerous and can it be prevented?

A.

Reverse runaway speed occurs when a pump runs in uncontrolled reverse direction due to the reverse flow of liquid from an elevated or pressurized source, such as a storage tank. It is most likely to happen with Vertical Turbine Pumps when pumping from deep wells.

A sudden power failure and/or discharge valve or non-return valve failure during operation against a static head may result in a flow reversal, and the pump will operate as a hydraulic turbine in a direction opposite to that of

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normal pump operation.

If the pump is driven by a prime mover offering little resistance while running backwards, the reverse speed may approach a maximum or runaway speed. If the runaway speed exceeds the normal pump speed, such speed may impose high mechanical stresses on the rotating parts of both pump and driver. Figure 2.58B on the right provides a means for estimating reverse runaway speed for pumps of different specific speed designs.

Vertical pump drivers can be equipped with non-reverse ratchets to prevent reverse rotation. However, their application is not always desirable and a review should always be made with the pump manufacturer. Check valves on the pump discharge and/or suction may also help with the problem.

Figure 2.58B â€” Reverse runaway speed ratio versus specific speed (US units)


Q. We operate about one thousand centrifugal pumps and find that ball bearing failure is a major cause of pump shutdown which requires repair, and sometimes causes unscheduled system shutdown. How can we monitor these pumps to eliminate or predict these failures?

A. Lubricant analysis is a good way to keep an eye on pump bearings and prevent

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unexpected bearing failure. If the lubricant is contaminated, it will accelerate the normal wear of bearings, and also show if excessive wear is taking place by the presence of metallic particles.

Quantitative analysis of metal particles are obtained from a particle counting analysis which will provide information on the number of particles in a sample within various size ranges and an analysis of the contaminants. A description of the methods and codes used can be found in ISO 4402 and SAE ARP 598.

Water is one of the most common contaminates found in large quantities and is the easiest to detect. Water can come from internal sources such as leaks in cooling systems and condensation, or external sources such as leaking seals. Usually the lubricant turns white and emulsifies. A sample from the bottom of the oil sump is best to indicate the presence of water.

Oil samples for the detection of particles should be taken while the pump is operating, or immediately after shutdown, to ensure well-mixed lubricant is obtained. Samples should be withdrawn from within one half inch of the surface of the oil. Until your experience provides statistics and consistency, it is suggested that samples be taken monthly.

For further detail on this subject, see ANSI/HI 9.6.5 Centrifugal and Vertical Pumps for Condition Monitoring.


Q. I know that centrifugal pumps do not operate well on viscous liquids and rotary pumps are often recommended, but what about reciprocating pumps? Are there any limits to the viscosity of liquids that can be handled by reciprocating pumps?

A. Theoretically there is no limit to the viscosity of liquids that can be handled by reciprocating pumps as long as they can be run slow enough.

The primary limiting factor is the velocity of liquid flow through the suction and discharge valves. This velocity can be calculated from the following equation:

v = 0.642Q/MA

Where

v = Average liquid velocity in the valve in feet per second;

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Q = Rate of flow in gallons per minute;

M = Number of suction valves;

A = Suction valve flow area in square inches.

Some guidelines in this regard would limit this velocity as follows:

Viscosity Valve velocity

SSU ft/sec

20,000 4.0

9,000 5.0

5,000 6.0

3,000 7.0

2,000 8.0

1,300 9.0

600 10.0

350 11.0

More information on this subject can be found in ANSI/HI 6.1-6.5 - Reciprocating Power Pumps.


Q. Our plant has a horizontal axially split case pump handling cooling water and the horizontal suction pipe includes a ninety degree elbow at the suction flange of the pump. The pump operation is noisy and one side of the impeller shows cavitation damage and the other side does not. I believe that the elbow on the suction is the major cause of the problem, but adding straight pipe in between is not practical. Is there another solution to this problem?

A. Yes, there are several solutions. One is to install one or more vertical flow dividers in the elbow which will guide the flow through the elbow and keep the flow to each side of the impeller equal. If the NPSH available is sufficient, the noise and cavitation damage should cease.

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Another solution is to reduce the flow velocity prior to the elbow with an eccentric increaser and add a reducing elbow at the pump. It may also be practical to use a combination of a larger elbow followed by a straight pipe reducer. The process of accelerating the flow rate has a tendency to straighten the flow at the same time.


Q. We are operating an end suction centrifugal pump which takes suction from the bottom of an open tank. As the level in the tank drops, a vortex is formed which allows air into the pump. This changes the performance of the pump and upsets the process downstream. I understand that this can be corrected with baffle plates. Are there any guidelines in this regard?

A. Yes, and baffle plates are effective in correcting this problem. The objective is to increase the apparent opening to the suction pipe, thereby reducing the liquid velocity as it leaves the tank. It is the velocity of the liquid flow that creates the vortex as the liquid level drops.

The simplest solution is the installation of a circular baffle plate with a diameter of four times the pipe diameter directly over the outlet pipe at a distance of two diameters above the pipe opening. See the figure on the right. This results in an apparent opening which is eight times the original opening and a corresponding reduction in flow velocity as the liquid moves under the baffle. Of course the velocity increases under the baffle plate, but then it is protected from the surface air.

If the baffle plate is supported by four to six vertical plates forming a star to a diameter of two times the pipe diameter, they will further reduce the start of a vortex.

For more information on this subject see ANSI/HI 9.8 - Pump Intake Design.

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Q. On low NPSH available applications it is often necessary to apply a centrifugal pump at 50% of its design flow rate to get a pump with low NPSH required values. What is the downside of doing this?

A. The most significant downside is in excessive power consumption. 50% of BEP is outside the normal operating range allowed by most pump manufacturers. This subject is discussed in detail in ANSI HI 9.6.3-1997 Centrifugal and Vertical Pumps for Allowable Operating Region.

Pump efficiency at 50% of BEP will always be lower than at BEP. If you do the math, you will find that the cost of excess power will be significant.

The lower NPSHR values at 50% flow do not guaranty cavitation damage free operation. At low rates of flow the impeller is also suffering from suction side recirculation in the impeller eye. The vortex caused by recirculation further increases local cavitation which in turn causes impeller damage, noise and vibration.

Another problem is the increase in radial forces on the impeller which reduce bearing and seal life and even cause shaft breakage from bending fatigue.


Q. In applications on handling slurries, we find that for higher heads centrifugal pumps are limited in what they can do. Can reciprocating pumps be used in these applications?

If so, are any special design changes necessary?

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A. Reciprocating pumps are often used to handle slurries for in-plant process and pipeline applications. Basic construction may be different from that for clear liquid applications. The differences may be in types of valves, addition of surge chambers, fluid injection into the inner portion of the stuffing box or material for wearing parts. ANSI-HI 6.1-6.5-2000 Reciprocating Power Pumps for Nomenclature, Definitions, Application and Operation contains considerable information on this subject.

Hydraulic passages should be sized so that the lowest velocity of the fluid will be above the critical carrying velocity of 4 to 6 ft/s. The highest velocity should be below that which causes excessive erosion.

Lubrication and flushing of packing are extremely important. Metered, clear, external injection, which is timed to the position of the plunger during its stroke, or continuous flow injection is required.

Valves for use in slurry service are designed for velocities between 6 and 12 ft/s to reduce erosion and abrasion of the valve seat and other valve components. Valve construction usually has replaceable valve inserts that are made of an elastomer or polymer. Metal to metal ball valves may also be used.

To facilitate starting and stopping a slurry pump, it should be fitted with adequate connections so the liquid end passages can be flushed of the slurry with clear liquid.

Rod and plunger packing requires special consideration when dealing with abrasive materials. In piston pumps, the piston runs in a renewable metal cylinder or liner. The liners are made of abrasion and corrosion resistant metals. Piston rods and plungers are coated to resist wear.


Q. Some centrifugal pump manufacturers show a vertical dotted line on their pump rating curves with the notation â€œminimum allowable flow.â€ � Is there a standard procedure for determining this limit? If not, what criteria are used to determine this value?

A. The factors which determine minimum allowable rate of flow include the following:

Temperature rise of the liquid -- This is usually established as 15Â°F and results in a very low limit. However, if a pump operates at shut off, it could overheat badly.

Radial hydraulic thrust on impellers -- This is most serious with single volute pumps and, even at flow rates as high as 50% of BEP could cause reduced bearing life, excessive shaft deflection, seal failures, impeller rubbing and shaft breakage.

Flow re-circulation in the pump impeller -- This can also occur below 50% of BEP causing noise, vibration, cavitation and mechanical damage.

Total head characteristic curve â€“ Some pump curves droop toward shut

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off, and some VTP curves show a dip in the curve. Operation in such regions should be avoided.

There is no standard which establishes precise limits for minimum flow in pumps, but ANSI/HI 9.6.3-1997 Centrifugal and Vertical Pumps â€“ Allowable Operating Region discusses all of the factors involved and provides recommendations for the â€œPreferred Operating Region.â€ �


Q. We are replacing a 30,000 gpm vertical pump with a larger one, and are concerned that the new pump will have enough submergence to prevent vortices at the pump inlet. What guidelines are available to determine if enough submergence is available?

A. 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 and are not addressed here.

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 + 0.574Q/D1.5

Where S is submergence in inches

D is bell diameter in inches

Q is rate of flow in gpm

The required minimum submergence can also be determined from figure 9.8.26B taken from ANSI HI 9.8-1998 Pump Intake Design.

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Q.

We operate a sanitary sewage pumping plant, which frequently creates vortices at the pump inlet pipes. The pumps are mounted in a dry pit adjacent to the collecting tank. The vortices are reducing the pump flow and causing excessive vibration in the pumps. Increasing the liquid level over the pumps would be very expensive. What else can be done to eliminate the vortices?

A. Plates, splitters, vanes, etc. can be added to the collecting tank near the pump inlet pipe. These will inhibit the formation of vortices. Assuming that the pump inlet pipe is in the wall between the collecting tank and pump pit, two suggestions could be tried.

Add a downward opening flanged elbow to the inlet pipe. This will lower the point where liquid enters and the flange on the elbow will reduce entrance velocity. If possible, use a reducing elbow to further increase the size of opening and flange diameter.

Attach a horizontal baffle plate to the wall directly above the pump inlet pipe. The plate should have a length and width equal to 4 times the diameter of the pump inlet pipe.

These solutions, and others, are discussed in a recently published 62 page standard, ANSI/HI 9.8-1998 Pump Intake Design addressing this subject. The standard provides design recommendations for a variety of different intakes for both solids bearing liquids and clear liquids.


Q. Our plant is operating pumps on cooling tower service which are experiencing cavitation damage to the cast iron impellers. The pumps are operating close to the best efficiency point and the NPSH available is slightly higher than the

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NPSH required by the pump. Should there be cavitation damage in the pump impellers under these conditions?

A. There is a common misconception among pump users that if the NPSH available from the system is greater than the NPSH required by the pump, that the pump will operate free of cavitation. This is not so. In order for a pump to operate cavitation free, the NPSHA must be from 2 to 20 times greater than the NPSHR of the pump. By definition, NPSHR is measured when the pump total head is reduced by 3% due to cavitation. For satisfactory operation, some NPSH margin over NPSHR must be provided by the system.

The Hydraulic Institute has recently published a new standard on this subject namely, ANSI/HI 9.6.1-1998 Centrifugal and Vertical Pumps for NPSH Margin. According to this standard, for cooling tower service, NPSHA should be 1.3 to 2.0 times NPSHR, depending on suction energy level, which is also defined in the standard.

The problem is also aggravated by the use of a cast iron impeller. In cooling tower service, other materials such as stainless steel, titanium and nickel aluminum bronze will withstand cavitation damage much better than cast iron.

Pump FAQsÂ® March 2006 Article #3

Q. We are using a large number of air-operated reciprocating pumps and are concerned about the amount of power that they are using. How can we check the efficiency of these pumps?

A. Air operated pumps generally do not show power or efficiency on their rating curves. See the typical curve on the right.

Note that instead of input power, lines of â€œconstant air consumptionâ€ are �shown. Air consumed by the pump is given in terms of the mass flow rate of the air through the pump at standard atmospheric conditions (68 degrees F and a pressure of one atmosphere). The standard unit of air consumption is SCFM (standard cubic feet per minute). Measurement of air consumption by the pump can be made and compared with the pump performance curve to check for deterioration. ANSI/HI 10.6 Air- Operated Pump Tests provides more detail on testing air-operated diaphragm and bellows pumps.

Reciprocating pumps generally have good hydraulic performance, which is well maintained. The power required by the compressor to produce the motive air is beyond the Hydraulic Institute scope.

Figure 10.13 - Plotting test results

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Q. What is a â€œcanâ€ pump?� I have seen this term applied to both centrifugal and vertical turbine pumps(VTP). Will you please explain it for me?

A. The term â€œcanâ€ pump may be used to describe either VTP, or close �coupled centrifugal pumps. However there is a big difference between them.

When applied to a close coupled centrifugal pump, the term â€œcanâ€ refers �to a cylinder or can which surrounds the motor rotor to seal it from the liquid being pumped. See figure 5.2, ref. number 221 on the right. In addition, another cylinder, ref. 217 is attached to the inner diameter of the stator, and is called the containment shell.

This design is more properly called a canned motor pump. See ANSI/HI 5.1-5.6, Sealless Centrifugal Pumps for more detail.

This â€œcanâ€ design is also used with rotary type pumps. When the term �â€œcanâ€ refers to vertical turbine pumps, the pump bowl assembly and �column pipe are inserted in a vertical â€œcanâ€ with sufficient clearance �between the â€œcanâ€ and bowl outside diameter to increase the �submergence of the first stage impeller without friction loss in the downward flow of the liquid.

Figure 5.2 â€” Canned motor pump: close coupled end suction, overhung impeller

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Figure 2.64 â€” First stage impeller datum closed suction â€“ can pump

See fig. 2.64 above and ANSI/HI 2.1-2.2, Vertical Pumps for Nomenclature and Definitions for more detail.


Q. I understand that pump vibration is an important parameter in predicting the imminent failure of a pump. Is this true? How is the vibration level measured, and what is an acceptable level of vibration for ASME B73 Chemical pumps?

A. Yes, vibration measurements are very useful in predicting pump failure. Such measurements should be taken on the outside of the pump bearing housing closest to the coupling. Measurements should be taken in both the horizontal and vertical direction perpendicular to the pump shaft with an instrument that measures in units of velocity in inches per second RMS, unfiltered. The graph from figure 9.6.4.4 shows the normally acceptable vibration level for ASME B73 pumps.

It must be understood that these are field vibration values for pumps in good

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condition, and exceeding them does not mean imminent failure. Your experience with pumps that have failed will help you determine an appropriate vibration level. More detailed information on this subject can be found in ANSI/HI 9.6.4, Centrifugal and Vertical Pumps for Vibration Measurement and Allowable Values.

Figure 9.6.4.4


Q. We operate a small plant and would like to improve the reliability of our centrifugal pumps and minimize unplanned pump shutdowns. Can you suggest a plan for monitoring pump performance to achieve our goal?

A. To begin with, we suggest a detailed review of each pumpâ€™s performance history to determine which are more likely to fail and why. Based on this review, take corrective action to upgrade the performance of the most troublesome pumps.

Monitoring the performance of all of your pumps is the next step. There are twelve main parameters which can be measured to prevent failure. These include:

1. Input power2. Temperature of key components such as bearings, seals, motor windings,

liquid pumped and more3. Corrosion of pressure containment parts4. Leakage from seals, gaskets, or pressure parts5. Pressure of liquid at pump suction and discharge6. Vibration at pump bearings7. Lubricant cleanliness or degradation8. Shaft runout at seal face9. Rate of flow of pumped liquid10. Maintenance inspection of critical components11. Pump operating speed12. Bearing wear

The frequency of recording these measurements depends on severity of the results of failure and the likelihood of failure of the weakest performers. See

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table 9.6.5.2 below for the recommended frequency measurement.

Implementation of this program requires considerably more information that is available from ANSI/HI 9.6.5, Centrifugal and Vertical Pumps for Condition Monitoring.


Q. I understand that the efficiency of variable speed drives is reduced as operating speed and power is reduced. Does this negate the power saving from the accompanying reduction in pump speed?

A. The difference in the power consumption with and without the variable speed drive must be evaluated. The power savings from the reduction in speed of the pump is typically greater than the loss in efficiency of the driver. Assuming a change in demand results in a reduction in pump speed of ten percent, the power required by a 100 horsepower pump is reduced by the cube of the speed so at ninety percent speed the power required is 0.90 cubed times 100 horsepower or 0.729 times 100 which equals 72.9 horsepower. At 90 percent of full speed the efficiency of the drive may be reduced by about ten percent of that at full speed and power. This depends on the design of the drive. If we assume the full load efficiency of the driver is 92%, then:

electrical power = pump power / driver efficiency

= 100 / 0.92

= 108.7 horsepower

At 90% of full speed:

electrical power = 72.9 / 0.92 X 0.90

= 88.0 horsepower


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Q. We are trying to standardize on a manufacturer of vertical pumps at our plant, but before doing so we are trying to figure out what is critical in the design of the pumps. In other words, what makes a vertical pump a GOOD vertical pump? The vertical pumps we use here are mainly in service all the time. We use them for pumping liquids from one tank to another part of the plant.

Any suggestions of where to look would be greatly appreciated.

A. I can appreciate your interest in wanting to standardize on a vertical pump manufacturer and understand the wisdom of finding a GOOD pump which is suitable for your applications.

Since you have indicated an interest in vertical pumps please refer ANSI/HI 2.1 â€“ 2.5 Vertical Pumps and ANSI/HI Vertical Pump Tests. These standards should help you to define what constitutes a GOOD pump design for your service requirements. You may also wish to look at ANSI/AWWA E-101 Vertical Turbine Pumps which was developed by the American Water Works Association. You may be interested to know that the HI Website [www.pump.org] will be up and running in mid-April. The website will include information regarding the HI, a supplier finder with manufacturer names cross-matched to specific products and hyperlinks to member websites.


Q. I take issue with the second question in FAQs in the February 2000 issue. You state that valves, elbows or other fittings attached directly to the pump discharge port will not effect what happens inside the pump impeller and casing. One of our customers installed a 500-hp double suction axial split case pump with horizontal discharge and a right angle elbow mounted horizontally on the discharge of the casing. Another right angle elbow was bolted to the first elbow. When operated near the best efficiency point, the pump bearing housings flexed vertically about Â¼ inch, and the pump rotor was shuttling back and forth axially. The pump bearings failed quickly as a consequence. When the discharge piping was changed to add ten diameters of straight pipe between the pump and the first elbow, the problem disappeared. Can you explain why?

A. We cannot provide a precise explanation of what took place inside this pump casing, but apparently the discharge elbows created a hydraulic feedback, which cause severe vibration in the impeller. The phenomenon is probably similar to that caused by suction side circulation in the impeller. We thank you for your input and ask if anyone else can explain what was happening.

Pump FAQsÂ® June 2006 Article #3

Q. When very large pumps, both high rate of flow and high horsepower are custom built, a smaller model may be built and tested to prove the proper

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performance of the design. What are the rules for building such a model?

A. The model impeller should not be less than twelve inches outside diameter. All dimensions of the model hydraulic passages, including the relative roughness of the hydraulic passages in the impeller and casing, should be in accordance with the model to prototype ratio.

The model test speed n1 and rate of flow Q1 shall be determined by the following relationship:

n1/n2 = (D2/D1)(H1/H2)0.5

Q1/Q2 = (D1/D2)2(H1/H2)0.5

Where D = diameter and H = head

More detail is available in ANSI/HI 1.6 Centrifugal Pump Tests.Pump FAQsÂ® June 2006 Article #2

Q.

I understand that NPSHR (required) by a pump is measured by the manufacturer using cold water, and I do not find information on how to correct this for other liquids. Hydrocarbons, for example, have a wide range of specific gravities and vapor pressures which must affect the NPSHR performance of a pump. Is this true?

A. Specific gravity and vapor pressure do not change the NPSHR of a centrifugal pump. By definition, NPSHR is determined by test and is selected as the value when the total head of the pump is reduced by three percent due to blockage of the flow through the impeller by the formation of vapor bubbles.

The most simple NPSHR test is conducted using a closed tank to which the pump is directly connected and the space above the liquid level in the tank is kept equal to the vapor pressure of the liquid. When testing with cold water, this is done by exhausting air or gas from the tank with a vacuum pump. Under these conditions, the NPSHA (available) is equal to the height of the liquid above the pump impeller.

When conducting the test, the rate of flow through the pump is kept constant and the liquid level in the tank is gradually lowered while total head measurements are recorded. The plot of test points will look like Figure 1.122 on the right.

If a liquid other than water is used, any difference in vapor pressure is compensated by a change on the surface of the liquid. In a similar manner, if a different specific weight of liquid is used, the reduced pressure due to the column height is compensated by the reduced weight of a given volume of

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pumped liquid.

There is another characteristic of liquid that does affect NPSHR. That is its thermodynamic properties. Some high vapor pressure liquids produce a lower volume of vapor when they boil and therefore less blockage in the impeller vanes. See ANSI/HI 1.3 Centrifugal Pumps for Design and Application for more information.

Figure 1.122 - NPSH Test With Rate of Flow Held Constant


Q. We use several metering pumps in our blending operation and are concerned about their accuracy. How accurate are these pumps and does accuracy change with different conditions?

A. Controlled volume metering pumps are reciprocating positive displacement pumps and are used in applications requiring highly accurate, repeatable and adjustable rates of flow. Accuracy is a compound definition composed of steady state accuracy, linearity and repeatability. Specific values for your pumps can be obtained from the manufacturer.

Steady state accuracy is the variation in the rate of flow over a specific period of time, under fixed pump and system conditions, expressed as a percent of the maximum calibrated rate of flow. Steady state accuracy applies over a defined turndown ratio.

Linearity is an expression of maximum deviation (plus or minus) of a series of measured rates of flow values versus rate of flow setting points to corresponding points on a theoretical best fit straight line drawn through the

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points on a graph. Linearity is expressed as percent of maximum calibrated rate of flow. See the figure on the right.

Repeatability is the rate of flow variation resulting from a specific excursion from a rate of flow set point, followed by a return to that set point, expressed as percent of maximum calibrated rate of flow.

More information on this subject can be found in the soon to be published standard, ANSI/HI 7.1-7.5 Controlled Volume Metering Pumps .


Q. I am aware that pump shafts must be carefully aligned with the driver shaft before start-up. This requirement is clearly stated in every pump manual. However, how carefully aligned is seldom answered. Is there a simple guideline for allowable misalignment measurements?

A. There is no single answer to this question. Misalignment often occurs in any of the following ways:

parallel offset angular offset combination of both

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axial movement

The extent of the misalignment in these modes depends on the tolerance of the driver, pump and coupling. By far, the coupling has the greatest ability to tolerate misalignment.

Elastomeric couplings can tolerate all forms of misalignment by the distortion of the elastomer. Most single metallic couplings such as gear, plate or grid type can tolerate only angular misalignment. To accommodate parallel offset, metallic couplings usually are supplied in pairs separated by a coupling spacer. The most forgiving is the universal joint when supplied as a pair with a spacer shaft. The allowable misalignment acceptable by the metallic couplings will depend on the length of the spacer. The coupling supplier can provide the allowable misalignment values for each type and size of coupling.

Parallel alignment can also be affected by thermal expansion during operation. Pumping hot liquids or even heat from the summer sun can affect the alignment. Pumps on hot applications should be realigned while the pump is at operating temperature.

Space between the shaft ends must also be provide to permit free axial movement of the shafts, especially with pumps on hot applications which will cause axial expansion of the pump shaft.

Pump FAQsÂ® July 2006 Article #2

Q. What is an API pump? I assume it is related to pumps used in the petroleum industry, but what is special about it?

A. As you guessed, an API pump is one that was built in accordance with standards published by the American Petroleum Institute. Five different pump types are described by the following standards:

API 610 Centrifugal Pumps for Petroleum, Petrochemical and Natural Gas Industries

API 674 Positive Displacement Pumps â€“ Reciprocating

API 675 Positive Displacement Pumps â€“ Controlled Volume

API 676 Positive Displacement Pumps â€“ Rotary

API 685 Sealless Centrifugal Pumps for Petroleum, Heavy Duty Chemical, and Gas Industry Services.

There are other API standards covering data sheets, sealing systems and sucker

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rods.

In common usage, the term API pump usually refers to an end suction centrifugal pump which is designed in conformance to API 610 for high temperature and pressure petroleum products. A primary feature is the support system which is at the centerline of the pump casing to minimize the effects of thermal expansion.


Q. During the maintenance of centrifugal pumps, how much wear of the wearing rings is considered normal, and what is the normal wearing ring clearance for a new pump?

A. A simple answer is not available. Normal wear of centrifugal pump clearance depends on many factors such as cleanliness of the liquid, viscosity of the liquid, the presence and size of abrasive solids, the head developed by the impeller and the operating speed.

The normal clearance range of a new pump can be obtained from the manufacturer or his representative, but the figure below can be used as a guide. Stainless steel impellers typically require greater clearance than bronze or iron to avoid galling and seizing.

(Insert figure 18 from the Student Workbook for the Energy Reduction Video - HERE.)


Q. We are considering the use of a magnetic drive pump for handling toxic liquids. The containment shell between the inner and outer magnet assemblies is very thin. We have concerns about wear in the shell and leakage of the liquid. Is our concern valid, and what can be done to protect against leakage?

A. Leakage of toxic liquid from any type of pump is a valid concern. You can take the following precautions regarding magnetic drive pumps:

The containment shell material must be selected to resist corrosion and erosion from the liquid being pumped. Magnetic containment shells shall be designed for the maximum allowable working pressure within the stress values for the materials in â€œSection VIII of the ASME Boiler and Vessel Codeâ€. � There shall be a minimum ratio between bursting pressure and design pressure of 2 to 1 for the pressure/temperature range in the pump.

The gasket between the containment shell and the pump cover casing shall be confined on the atmospheric side to prevent blowout. The design shall consider thermal cycling which may occur as a condition of service.

Sensors should be provided to measure and monitor the position of the inner and outer magnet assemblies to prevent them from rubbing on the containment shell.

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A secondary containment shell shall also be provided. This is usually done by the outer housing which surrounds the outer magnet assembly, and a suitable sealing device around the input shaft on the pump side of the inboard bearing (Ref 16). See the figure below:

Provide a leak detector or pressure sensor at an appropriate location in the outer containment shell (frame, part 19 or cover, casing, part 239) to detect loss of primary containment.

For further information see ANSI/HI 5.1-5.6-2000 Sealless Centrifugal Pumps.

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Q. I have been told that the performance characteristics of centrifugal pumps are changed when pumping viscous liquids such as hydrocarbons. How can this performance change be calculated?

A. The answer to your question is not that simple. The members of the Hydraulic Institute and a team of international experts put their collective experience

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together and prepared a thirty-page document on this subject which was approved as an American National Standard. It is designated as ANSI/HI 9.6.7 Effects of Liquid Viscosity on Rotodynamic (Centrifugal and Vertical) Pump Performance.

The correction factors used in this standard are based on in-depth hydraulic design calculations proven by extensive laboratory test data of many pumps on liquids with a wide range of viscosities.

The correction factors are used to adjust the performance on water as shown below in Figure 9.6.7.1 a and b.

Figure 9.6.7.1 â€” Modification of pump characteristics when pumping viscous liquids


Q. We recently experienced line shaft failure in a multistage vertical turbine pump with ten-inch diameter bowls. An analysis of the break pointed to torsional fatigue as the cause but the pump was operating under full load in a steady mode and vibration monitors did not show excessive vibration. How can the shaft fail in fatigue under these conditions?

A. The existence of torsional vibration rarely shows itself in the pump column or driver housing vibration. Thus, pump shafts, coupling or gears can fail without the usual (radial) vibration monitoring equipment indicating any danger.

Rotor torsional critical speeds can be present any time there are two rotating masses connected by a shaft that is not infinitely stiff. This implies that any pump rotor coupled to a driver has a torsional natural frequency of vibration. Torsional critical speeds are associated with torsional or angular deflection of the rotor and are not to be confused with lateral critical speeds associated with lateral deflection. The resulting stresses and angular deflections can cause

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premature equipment failure.

With vertical pumps, the primary hydraulic exciting force is generated by the impeller vanes passing the bowl vanes. The energy level of this force is typically low, with resulting negligible vibration amplitudes.

On the other hand, when a pump is driven through a gear, the inaccuracies in the gear teeth can provide the exciting force and high torsional critical speeds at tooth meshing frequencies. Similarly, engines can also cause high torsional vibrations at higher frequencies.

For more information on pump vibration, see ANSI/HI 9.6.4 Centrifugal and Vertical Pumps for Vibration Measurements and Allowable Values.


Q. When checking the performance of a new centrifugal pump, the plot of the resulting head versus rate of flow curve appears to be lower than the manufacturerâ€™s rating curve. How much deviation from the manufacturerâ€™s rating curve is normal?

A. The Hydraulic Institute Standard contains two performance test acceptance tolerance levels, "A" or "B" which must be agreed to by both customer and manufacturer. The acceptance tolerance applies to the specified condition point only, not to the entire performance curve. It is recommended that the contractual agreement contain the agreed upon acceptance level. The tolerance for total head has four different categories that depend on the total head and rate of flow. There is an alternate tolerance for rate of flow at the rated total head. All of this is dependent on an agreement with the pump manufacturer to perform a factory test to determine the true performance and take any necessary corrective action before the pump is shipped. Field tests are seldom accepted as reliable due to the difficulty of meeting the proper test procedures in the field and making corrections.

Assuming acceptance level â€œAâ€ is applicable and the rated conditions are �3500 gpm at 300 feet, the head variation at the rated flow is +5% to -0%. An alternate tolerance in rate of flow at rated head is +10% to -0%.

Additional details are available in ANSI/HI 1.6 Centrifugal Pump Tests.


Q. When installing a horizontal pump, is it better to grout the pump baseplate or let it float free? Different sources provide opposing views on this subject.

A. Whether to grout or not depends on the specific application and the design of

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the baseplate. Applications that undergo wide temperature variation may do well with floating baseplates in order to allow for movement of the pump and thereby minimize pipe strain caused by thermal expansion of the pipe. However, the baseplate in such cases must be sufficiently rigid and designed to avoid misalignment of the shaft coupling as the baseplate moves.

Functional requirements:

The purpose of a baseplate is to provide a foundation under a pump and its driver that maintains shaft alignment between the two. This baseplate must allow for initial mounting and alignment of equipment, survive handling during transportation to the installation site, be capable of being installed properly with minimum difficulty, allow final alignment of the mounted equipment, and allow removal and reinstallation of equipment. It must be recognized that it is not necessary that an absolutely rigid baseplate be designed to meet these requirements. At the same time, the baseplate must not be permanently deformed after the equipment is mounted at the manufacturing facility. Compliance with these design criteria, in conjunction with proper installation procedure, will contribute significantly to meeting the functional requirements.

Free standing baseplate:

A free standing baseplate is a design which is intended to be elevated off the floor or deck and supported by stilts or shims. This type of baseplate must be designed to provide its own rigidity as there is no grout for support. See Figure 1.94.

Additional information and calculation of baseplate stiffness can be found in ANSI/HI 1.3-2000 Centrifugal Pumps for Design and Application


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Q. We are designing a circular sump for the installation of three sewage pumps. Because of local restrictions on the size, we are concerned that the pump intake structure may not provide a good hydraulic design and are considering whether to build and test of a model of the intake. How do we select the appropriate rate of flow for the model test?

A. Models involving a free surface are operated using Froude similarity since the flow process is controlled by gravity and inertial forces. The Froude number, representing the ratio of inertial to gravitational forces, can be defined for pump intakes as:

F = u/(gL)0.5

Where:

u = average axial velocity (such as in the suction bell entrance) in ft/sec

g = gravitational acceleration, 32.2 ft/sec2

L = a characteristic length (usually bell diameter or submergence) in ft.

The choice of the parameter that is used for velocity and length is not critical, but the same parameter must be used for the model and prototype when determining the Froude number. For similarity of flow patterns, the Froude number shall be equal in both the model and prototype and solving for the velocity in the model will answer your question.

In modeling a pump intake to study the potential formation of vortices, it is important to select a reasonably large geometric scale to minimize viscous and surface tension scale effects, and to reproduce the flow pattern in the vicinity of the intake. In addition, the model shall be large enough to allow visual observations of flow patterns, accurate measurements of swirl and velocity distribution, and sufficient dimensional control. Realizing that larger models, though more accurate and reliable, are more expensive, a balancing of these factors is used in selecting a reasonable model scale. However, the scale selection based on vortex similitude considerations is a requirement to avoid scale effects and unreliable test results. Fluid motions involving vortex formation have been studied by several investigators (Anwar, H.O. et al., 1978; Hecker, G.E., 1981; Padmanabhan, M. and Hecker, G.E., 1984; Knauss, J., 1987)

ANSI/HI 9.8 Pump Intake Design includes additional information on this subject.


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Q. We operate a vertical turbine type well pump that has a clutch at the top of the motor that often disengages when the pump starts. What causes this to happen and what can be done to prevent it?

A. Vertical turbine pumps are designed so that the downward thrust due to the weight of the rotor and hydraulic force from pressure on the back of the impellers is carried by a high thrust tapered roller bearing on the top of the motor. This bearing is not capable of carrying any thrust in the upward direction. However, during pump operation there is also an upward thrust on the pump rotor caused by the change in momentum of the flow as it changes direction from axial to radial at the entrance to the impeller blades.

This up thrust is usually less than the heavy down thrust from the weight of the rotor and hydraulic forces so that no net up thrust occurs. However, at the moment of start up, there is no resistance to flow and the pump operates momentarily at a very high rate of flow. In this case there is little down thrust from the impeller and maximum up thrust from the change in momentum resulting in a net up thrust which disengages the motor clutch.

To correct this problem the pump rotor needs to be fitted with a thrust collar, which is capable of resisting the momentary up thrust.


Q. Normal pump operation will eventually corrode and erode the inside of a pump casing until the thickness of the casing wall is too thin for safe operation. How can this safe thickness be determined and what are the potential consequences of operating with less than proper thickness?

A. The minimum casing thickness for a pump is determined by the manufacturer based on the maximum casing working pressure and temperature and the material of the casing. This is done using finite element analysis or empirical calculations based on test. The ASME Pressure Vessel Code is used to determine the proper factor of safety for the casing material and the manufacturer publishes the resulting minimum thickness. Pump manufacturers also add a corrosion allowance to the minimum wall thickness so the casing as manufactured can safely be eroded by the amount of the allowance.

Operating a pump with less than minimum wall thickness should be avoided. However, the pump will not immediately fail because of the factor of safety and possible operation below the maximum working pressure. The consequences of failure will range from a leak in the casing wall to total fracture.

Pump FAQsÂ® April 2002 Article #2

Q. Our plant uses a 10-inch vertical turbine pump, which is about 30 feet long, to

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supply cooling water. The requirements have been changed and we need an increase in flow of about 10%.

According to the pump rating curve, the pump should satisfy our needs, but it does not. We have renewed the impeller wearing ring clearances, but we cannot reach the rating curve performance. What else can be done? We cannot increase the impeller diameter due to power limitations.

A.

Increasing the pump column pipe diameter will certainly help. Most pump manufacturerâ€™s performance curves are limited to the bowl assembly performance, since the length and diameter of the column pipe represent an unknown variable.


Q. We have difficulty in getting wearing ring data for several older pumps. Is there a general guideline on recommended wearing ring clearance that we can use?

A. In 1997 the Hydraulic Institute published a listing of recommended clearances for a range of wearing ring diameters. This was included in an educational video titled â€œEnergy Reduction in Pumps and Pumping Systems.â€ �


Q. I understand the importance of monitoring pump performance in order to prevent unexpected pump failures. Performance parameters such as discharge pressure, rate of flow, input power, and vibration are straight forward. But how do we know when corrosion or erosion have reduced pump casing wall thickness to the point at which casing failure may occur?

A. The casing wall thickness must be measured periodically to know when failure is imminent. But you know this. Measuring the casing wall thickness is the problem. To this end, several methods can be used:

Visual inspection is the easiest and may be the most economical method. However it is cumbersome and messy when dealing with corrosive liquids. In addition, stress corrosion can occur without any visible signs, resulting in sudden failure.

Electrical resistance measurement of the casing wall using a metal probe shows an increase in resistance as the casing thickness is reduced by corrosion. However this is not useful in detecting localized forms of corrosion such as pitting.

Linear polarization is another method that involves measurement of a current response to an applied potential through probes that are inserted in the system. To use this method a conductive liquid is required.

Ultrasonic thickness measurement can also be used. To use this method the casing surface must be cleaned to bare metal. Ultra sound is not as accurate as the other methods.

The frequency of wall thickness measurement depends on the expected

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corrosion rate of the pump on the application. The last three methods lend themselves to continuous monitoring or frequent checks by use of small, portable devices. The visual method is more difficult, but the first measurement should be made no longer than 3 months after start up.

Pump casings are designed with a built in allowance for corrosion. This allowance can be supplied by the pump manufacturer. Contact the pump manufacturer or his representative when the allowance has been reduced by 50% or if a significant change in process conditions causes rapid reduction in wall thickness. According to ANSI/HI 9.6.5 Centrifugal and Vertical Pumps for Condition Monitoring, when measurements show that 70% of this corrosion allowance has been lost, the pump should be â€œShutdownâ€.�


Q. Pumps operating in wet pit applications often suffer from the accumulation of grit or other solid material which accumulates near the pump inlet. This requires draining of the pit or removal of the pump to remove the solids. Is there a better way?

A. It is often practical to build the pit with sloping sides to minimize the horizontal floor area. See the figure for an example.

The sloped walls should be a minimum of 60 degrees from the horizontal for concrete or 45 degrees for steel. This design allows the solids to collect at the pump inlet where they can be swept away with the flow. The pump should be periodically operated at its maximum rate of flow to more effectively carry away the solids. If two or more pumps are in the pit, the one farthest from the inlet to the pit may be run alone to remove the solids. Verify the NPSHA to ensure there is sufficient margin to operate at the high rates of flow. ANSI/HI 9.8 Pump Intake Design includes considerable detail on this design.


Q. When pumping liquids with entrained solids, especially slurries, I understand that hard material for the impeller and casing is recommended to resist erosion. I also hear that soft rubbery material is also used. How can this be? When are rubbery materials best used?

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A. As a general guideline, hard metals are often used in applications characterized by large, sharp edged solids, and elastomers for smaller round-edged solids. Either high- chrome irons or elastomers are used for their corrosion resistance. In special applications with low head requirements, solid ceramic-lined pumps are used for pumpages containing fine material.

Pump speed is also a factor. Impeller tip speed, as distinct from rotational speed, is often used as a guide for wear in the selection of slurry pumps. ANSI/HI 1.3 Centrifugal Pumps for Design and Application recommends the following maximum impeller tip speeds:

dirty water-130 ft/sec medium slurries up to 25% solids and 200 micron solids- 115 ft/sec higher slurry concentrations and larger solids-100 ft/sec

pumps fitted with elastomer impellers- 85 ft/sec.

Pump FAQsÂ® April 2005 Article # 1

Q. We operate a number of end suction centrifugal pumps on a chemical process application. The seals operate well, but the anti-friction bearings fail in less than twelve months. Is this normal? If not, what can be done to increase bearing life?

A. No, pump bearings are usually selected to provide a minimum of two years continuous operation before failure and the average bearing life is about five times longer. The next time the bearings fail, check for the following during the bearing replacement process:

Shaft coupling alignment: Poor alignment imposes additional loads on the bearings which will reduce bearing life. For pumps operating in higher temperature service, hot alignment checks are recommended. Shaft alignment can be affected by nozzle loads. Be sure that when the pump is running, the nozzle loads are within acceptable levels. Normally, however, poor alignment will also lead to premature seal failure.

Lubricant amount, condition, and cleanliness: Corrosive chemicals, water, or solids in the oil will attack the bearings and reduce life. Make sure the bearing cover seals are replaced each time, and if necessary retrofit the pumps with more elaborate labyrinth bearing seals. Also make sure that the shaft-mounted rotating flingers are close to the bearing seals to help keep contaminants away. Follow the manufacturerâ€™s recommended lubrication intervals. Where grease lubrication is used, use caution and do not overlubricate the bearing.

Bearing fit on the shaft: If the shaft diameter under the bearing is too large it will expand the ID of the inner race of the bearing excessively and preload the rotating elements of the bearing beyond the recommended levels. Check the bearing manufacturerâ€™s catalog for the recommended shaft dimension. Verify that the internal preload of the bearing meets the recommendation of the pump manufacturer.

Bearing assembly procedure: When the replacement bearing is pressed onto the shaft, make sure that the pressing force is applied evenly to the inner race only and not to the outer race. Be sure to follow the manufacturerâ€™s recommendations when preheating the bearing is required. Care shall be taken to install the bearing square with the pump shaft, providing uniform contact with all shoulders, rings, and checknuts on

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the shaft and in the bearing housing. Coupling half assembly: Most couplings are designed with a light

interference fit on the shaft. If the coupling half is assembled to the shaft with a hammer or press, the restraining force should be applied to the impeller end of the shaft and not to the pump frame, which will transfer the load to the bearings.

Finally, when back in operation check that the pump is operating within the manufacturerâ€™s allowable operating region. Verify that the bearing temperature and vibration levels are within acceptable limits.


Q. I have been told that reciprocating pumps must be protected against operation with a closed discharge valve, but that is not necessary with air operated diaphragm pumps. Is this true and why?

A. Yes, it is true. Motor driven reciprocating pumps continue to rotate when the discharge valve is closed. The closed valve prevents the fluid from exiting the pump, subjecting the motor and other pump components to excess internal stress due to the high torque that is generated by the motor as load increases and speed is reduced by that load. Bypass valves, clutches and other mechanisms are necessary to allow safe operation in this condition. With air operated pumps, when the discharge valve is closed, the air and fluid pressures equalize and the diaphragm or bellows stops reciprocating. There is little additional stress generated because the pressures on both sides of the diaphragm are equal. The pump end is designed to withstand the force generated by the maximum rated air pressure so as long as the air pressure is below the rated air pressure, the pump cannot be damaged. The Hydraulic Institute standard, ANSI/HI 10.1-10.5 Air Operated Pumps contains much useful information on these pumps.


Q. The December 2004 issue of P&S included a question and answer on air entrainment in rotary pumps which I found to be useful. Is there similar information on rotary pumps handling liquid with solids or slurries?

A. Yes, ANSI/HI 3.1-3.5 Rotary Pumps includes several good pages on this subject.

Rotary pumps may be used for in-plant process and pipeline transfer of slurries when metered flow or medium-to-high discharge pressures are required. Since volumetric efficiency and therefore mechanical efficiency are normally dependent on the clearances between the pumping elements of a rotary pump, care must be taken in the selection and application of the pump in slurry service. Slurries containing hard particles can cause abrasive wear in rotary pumps.

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The size and general shape of the solids in a slurry normally govern whether a particular rotary pump configuration and/or clearances can be used for the slurry in question. Usually the clearances must be greater than particle size. In some pumps, the rotor configuration will accept large particles as long as the size, distribution and shape are controlled. Thus, the size must be related to porting, rotor cavity size, configuration and interaction and operating clearance.

The effect on rotary pump performance can vary widely as slurries change with time, control, character, etc. Low concentrations of fine non-settling solids in a Newtonian fluid carrier may have no appreciable effect on either the power requirements or the pump rate of flow. Generally, as the percentage and size of the solids increase at given conditions of operation, speed and pressure, the pump input power curve increases (see Figure 3.34).

Although rotary pumps are capable of limited slurry handling, the particular rotor and housing configuration make the various rotary designs more or less adaptable to specific types of slurry. Many slurries similar to paper stock require open porting and clearances and definite minimum velocities of flow. Clay slurries require low shear rate.

Figure 3.34 - Differential Pressure Vs. Pump Input Power


Q. What is a balancing disc or drum? I understand that multistage pumps, such as on boiler feed service, usually use such a device to balance hydraulic thrust.

Figure 1.21 below shows the last two stages of a multistage pump with a balancing disc shown on the left. The space between the last impeller and the balancing drum is under full discharge pressure which creates a force to the

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right over the area of the impeller eye, which at the suction end is under suction pressure. The resulting force can be very large, and difficult to restrain with a thrust bearing.

A. To counterbalance this force a disc or cylindrical drum, with the same diameter as the impeller wearing ring, is mounted on the shaft. The hydraulic pressure acting on this device creates a force in the opposite direction from the impeller thrust, thereby balancing the force on the impellers. This reduces the net axial unbalance to practically zero.

Figure 1.21 shows a balancing disc which is slightly larger than the wearing ring. The pressure here will push the pump rotor to the left, but as the disc moves, the space between the disc and stationary member opens with a resulting flow of liquid which is pipe back to the suction.

There is a close clearance bushing between the impeller and disc so as liquid flows the pressure on the disc drops and the space between the disc and stationary member closes. This makes the rotor self adjusting so the net thrust is always zero.

If a cylindrical drum is used in place of the disc, the thrust can be effectively balanced, but the self adjusting feature will be lost, and an appropriate thrust bearing will be required.

Figure 1.21 - Multistage Pump with Balancing Disc

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Q. We need to replace several pumps handling paper stock, and need to know what special problems we should look out for. Can you help with such information?

A. Paper stock varies considerable depending on the wood source that is used and the consistency of the stock.

Low-consistency stock usually refers to a class of products with 1 to 7 percent fiber content by weight. These paper stocks are normally handled by end-suction centrifugal pumps equipped with semi-open impellers and contoured wear plates.

Medium-consistency stocks are made of 8 to 15 percent paper fiber. The

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rheological properties of fiber-water suspensions in this range are dependent on the properties of the individual fibers and the viscoelastic network that they form. Special designs of centrifugal pumps are required to handle this type of paper stock. For example, some form of â€œshear generatorâ€ is needed at the �inlet to create turbulence and reduce the effective fluid viscosity. Special impeller design and air-extraction devises are also required to prevent air binding.

An end-suction centrifugal pumping unit must be specifically designed to handle medium consistency stock mixtures without clogging the device, or dewatering the stock. A large suction-eye and unobstructed waterways can be provided by an overhung, semi-open impeller design. This keeps the suction velocity low to promote smooth flow, avoid air binding and prevent separation of stock fibers from water. The contoured front surfaces of the impeller vanes interface with the replaceable wear plate. This arrangement provides a self-cleaning effect whereby the impeller resists clogging to improve its reliability.

High consistency paper stocks contain more than 15 percent paper fiber, and are found in the bleaching operation. Centrifugal pumps cannot handle such high consistency, and therefore positive-displacement rotary units are used. Proper suction piping design has to be included to help this high solids mixture to enter the suction cavities of the rotary pump.

More information on this subject can be found in ANSI/HI 1.3 Centrifugal Pumps for Design and Application.

Pump FAQsÂ® August 2004 Article #2

Q. Our company operates a high pressure process during which the liquid is throttled to low pressure after processing. We are considering the use of a power recovery turbine to recover some of the energy lost by throttling. Can a typical centrifugal pump be used as a turbine, and if so, what precautions must be taken?

A. Yes. A typical centrifugal pump can be used as a turbine, however, selection and installation must be done carefully. While operating in the turbine mode, the performance characteristics of a pump as turbine (PAT) differ significantly from operation in the pump mode.

The PAT should be selected by the pump manufacturer to insure that the selection is both hydraulically and mechanically suited for the application. Precautions must also be taken to insure that the PAT will operate without cavitation. The turbine industry uses the terminology TREH (Total Required Exhaust Head) and TAEH (Total Available Exhaust Head) in place of NPSH.

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Some of the other factors which affect the use of pumps as turbines include:

Runaway speed Liquid flow at runaway speed Required solids passage Liquid borne abrasives Torque reversals during start-up or shut-down

Overspeed trip and control


Q. One pumping system in our plant makes a load bang when it is shut down. I have been told that this results from water hammer. What causes water hammer and what can be done to correct it?

A. Water hammer may have severe effects and damage parts of the pumping system. It results when liquid flowing through a relatively long length of pipe is suddenly stopped. The velocity energy of this large mass of liquid is suddenly converted to pressure energy with a resounding bang. Itâ€™s like trying to stop a long railroad train within in a short distance.

Some suggestions for correction include the following:

Reduce the liquid flow gradually before stopping the pump Add a liquid chamber with an air cushion to absorb the energy Reduce the pump speed slowly by using a variable speed drive or by

adding a flywheel

Install a check valve that closes slowly to minimize shock


Q. When purchasing pumps for water supply systems, is there a simple way to determine the maximum allowable speed for a specific application? If so, what factors should be considered?

A. The primary factor to be considered when estimating the maximum pump speed for water supply systems is minimizing the potential for cavitation damage. Experience has shown that cavitation risk is minimized when the pump suction specific speed is equal to or lower than 8500 in US units. Suction specific speed is a characteristic of a pump, which is determined by the following equation:

S = nQ0.5 / NPSHR0.75, where:

S = suction specific speed

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n = speed of pump in rpm

Q = pump rate of flow in gpm for single suction pumps and half total flow for double suction pumps

NPSHR = net positive suction head required in feet

If we solve this equation for n, substitute 8500 for S and NPSHA for NPSHR, we get:

n = ((8500) x (NPSHA)0.75) / Q0.5

Substituting the systems values for NPSHA and Q will give the maximum recommended speed for this application. Q can also be varied by changing the number of simultaneously operating pumps in the system

Another consideration is maximizing pump efficiency. Again, experience has shown that centrifugal pump efficiency is maximum when pump specific speed is between 2000 and 4000 in US units. Specific speed is determined by the following equation:

NS = nQ0.5 / H0.75 where:

NS = specific spee

n = speed of pump in rpm

Q = pump discharge rate of flow in gpm for both single suction and double suction pumps

H = pump total head in feet.

The figure below shows the efficiency correction or reduction as a function of specific speed. Refer to ANSI/HI 1.3 Centrifugal and Vertical Pumps for Design and Application for additional details.

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Q. I understand that the radial force on a centrifugal pump impeller increases as the pump is throttled to lower rates of flow. Why does this occur and how can this force be calculated?

A. In a properly designed centrifugal pump, the distribution of flow and pressure around the impeller is uniform at the design rate of flow. Consequently, there is little if any unbalance of forces around the impeller and very low net radial force. When the rate of flow is reduced by throttling, some of the liquid from the impeller is forced back into the casing volute causing an unbalance of flow and an unbalance in pressure distribution. The flow velocity is higher in the vicinity of the volute tongue and the pressure is lowest at that point. The result of this is an unbalance in pressure distribution around the impeller and a net force or thrust in the radial direction.

This phenomenon was studied experimentally in the early 1950s, and the following equation was developed:

Where: RT = radial thrust in pounds

K = thrust factor from figure 1.81 below

H = pump head per stage in feet

s = specific gravity of the liquid

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D2 = impeller diameter in inches

b2 = impeller width at discharge including shrouds in inches

The radial thrust must be supported by the pump shaft and bearings. Higher radial thrust loads may contribute to shorter bearing life and deflection at the pump seals causing premature seal failure. More detail on this subject can be found in the American Standard â€œANSI/HI 1.3 Centrifugal Pumps for Design and Applicationâ€�.


Q. How can we get longer life from the ball bearings in our pump? Our process uses an end suction centrifugal pump to circulate heat transfer liquid at 750 degrees F. The bearing housing is cooled, and the bearings look clean after they fail. What else should we look for?

A. Pump bearings are most likely to fail due to contamination and poor lubrication. However, that does not seem to be the case since the bearings are clean. In addition, the bearing cooling is a big help.

That leaves the matter of excess bearing load. This is also consistent with the regularity of the failures.

One common cause of high bearing load is coupling misalignment. If the coupling halves are aligned when the pump is cold, they will usually move out of alignment as the pump heats up. Even with the pump supports at the shaft centerline, some movement from pipe stains is still possible. To avoid this problem, shut down the pump when it reaches operating temperature and recheck the coupling alignment. Realign the coupling if necessary.

Another cause of high bearing load is operation below fifty percent of the pumps best efficiency rate of flow. At low flow rates, all of the flow from the

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impeller cannot exit from the pump casing so it recirculates in the impeller. This causes an uneven distribution of pressure around the impeller resulting in a high radial force on the impeller which must be supported by the bearings. To guard against this, determine the actual rate of flow through the pump and make sure it is close to the pumps best efficiency flow.

Another problem which is unique to high temperature applications is excessive expansion on the bearing inner race. Even with bearing cooling, heat is transmitted to the bearing through the pump shaft. This heat causes the shaft diameter and the inner raceway of the bearing to expand which in turn squeezes the rotating balls in the bearing. The result is shorter bearing life which can be corrected by using bearings with greater internal clearance. Such bearings may also be referred to as C3 Fit bearings.


Q. The hydraulic coverage charts for standard pump lines published by manufacturers form a rectangular grid with pump size varying with rate of flow horizontally and total head vertically. How do pump manufacturers determine the design rate of flow and total head for each pump in the series?

A. The horizontal spacing between pumps is determined by a reasonable flow velocity for the discharge opening of the pumps. The popular pipe sizes, 1, 1Â½, 2, 3, 4. 6 inch etc. also double in area in progression. Hence, pumps increase in rate of flow in a doubling progression.

Total head between pumps usually increases by a factor of 1.6. This allows an acceptable cut down range between pumps without an excessive reduction in efficiency at the lower end of the pumps coverage. These factors of 2 for rate of flow and 1.6 for head also result in diagonally spaced pumps having the same design specific speed (NS). For example, a 3 inch discharge pump with a 10 inch impeller has the same NS as a 4 inch with 13 inch impeller and is therefore geometrically similar. This makes the designerâ€™s job easier.

Finally, experience, competition and standards play a big role as each manufacture attempts to compete with all others.


Q. Many sources, including the Hydraulic Institute e-learning program Centrifugal Pumps: Fundamentals, Design and Application, describe the use of Affinity Laws which are used to predict the performance of a centrifugal pump when the impeller diameter is reduced or cut down. One of the laws says that Q1/Q2 = D1/D2, where Q is the rate of flow and D is the impeller diameter. However, ANSI/HI 1.6-2000 Centrifugal Pump Tests says

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that at a given speed, Q1/Q2 = (D1/D2)2 (H1/H2)0.5 which is equivalent to Q1/Q2 = (D1/D2)3. Which is correct, or is there some explanation for the difference?

A. Both are correct for their intended use. The Affinity Laws are used to predict the performance of an existing impeller of known performance when its diameter is cut or reduced in the order five or ten percent. This is an empirical relationship and is reasonably accurate for diameter reductions up to five percent, and to some degree up to ten percent.

The modeling laws as described in the HI Test Standard are used to compare the performance of two distinct pumps which are of different size but geometrically similar.

They are also used by pump designers when they design a new pump based on the geometry of an existing pump of a different size. The modeling laws are very accurate provided that the surface roughness of the waterways as well as all other dimensions of the impeller are kept to the same proportions as the original. In addition to the impeller dimensions the casing dimensions are modeled as well. This is a true three dimensional comparison, so the rate of flow changes as the third power of the impeller diameter.

On the other hand, when the diameter of an existing impeller is cut down, no other dimensions change, and the effect on rate of flow is linear.


Q. I understand that pipe strains are bad for pumps and should be avoided. However, some installations inherently have this problem due to temperature changes which result in thermal expansion of the piping. Expansion joints are designed to relieve such forces, but what else can be done to solve this problem?

A. Expansion joints have a problem too. Typical expansion joints do not restrain the axial forces in the piping due to the liquid pressure. For example, the axial force caused by 150 psi pressure in a six inch pipe is 4,241 pounds. This force is usually restrained by the pipe, but with an expansion joint, the pipe do not do this. Such forces must be carried by the pump if the expansion joint is close to the pump. The maximum allowable force on the discharge flange of an 8X6X13 ANSI B73.1 pump is 3500 pounds according to ANSI/HI 9.6.3-2000 Centrifugal and Vertical Pumps for Allowable Nozzle Loads. At 150 psi, this allowable limit would be exceeded.

Another approach to this problem is a free floating pump or a spring mounted baseplate. Close coupled pump designs can facilitate this approach. Vertical-in-Line pumps are a variation of the close coupled design and are available for many process applications. VIL pumps can transmit substantial forces through the casing between the discharge and suction pipe. ANSI/HI 9.6.3 shows and

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allowable force between discharge and suction of a 4 inch discharge pump to be 18,704 pounds.

Spring mounted baseplates can also work well, but the baseplate must have sufficient stiffness to maintain coupling alignment without relying on a concrete foundation.


Q. I understand that centrifugal pumps are not well suited to handle viscous liquids. Is there an easy way to evaluate the viscosity value above which centrifugal pumps are not recommended?

A.

Centrifugal pump performance is significantly impaired when pumping liquids with viscosities that exceed 1000 SSU. Above 10,000 SSU they are almost useless. Additional guidance on this subject may be found in the Hydraulic Institute Standard ANSI/HI 1.3â€”2000 Centrifugal Pumps for Design and Application. Page 25 of this standard contains the correction chart that appears in Figure 2 below.

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Q. We are operating an end suction pump with an elbow connected to the suction

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flange. The suction connection is eroded on one side due to the uneven flow coming out of the elbow. Is there a simple solution?

A. An improvement can be achieved by using a reducing elbow at the pump. Reducing two pipe sizes is better than one. If the existing suction line will not permit this, it should be increased to 2 sizes larger than the connection at the pump.

A few years ago, one pump manufacturer designed a special elbow that looked similar to the one shown in Figure 1. Tests showed that this design provided a more uniform flow output.


Q. One of our utility pumps takes suction from a river that frequently becomes muddy. The bronze impeller and wearing rings become badly eroded more quickly than anticipated. What do you recommend to increase useful impeller life?

A. There are several approaches you may want to consider: 1. Reduce the solids content from the liquid by installing a settling chamber or

cyclone separator upstream of the pump suction.2. Operate the pump close to its best efficiency point (BEP). This will help to

optimize the angle with which the vane and liquid meet thereby minimize erosion.

3. Upgrade the impeller material to something more abrasion resistant such as series 400 stainless steel.

4. Replace the pump with one that is specifically designed for abrasive

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service. You can locate abrasive pump suppliers on

www.pumps.org under the Supplier Finder section.


Q. Is it an acceptable practice to select centrifugal pumps at rates of flow that are less than 50% of the Best Efficiency Point (BEP) in order to get a pump, which requires lower values of Net Positive Suction Head Required (NPSHR)? Most pump curves indicate lower values of NPSHR at lower rates of flow.

A. Continuous operation of centrifugal or vertical pumps below 70% of the BEP may result in the following problems:

Recirculation in flow occurs on both sides of the impeller, which causes cavitation damage to the impeller, excessive noise, and high vibration of the pump rotor.

Unbalanced radial forces, which act on the outside diameter of the impeller, increase as the rate of flow is reduced. This phenomenon is typically more severe in single volute designs. It results in reduced bearing and mechanical seal life and may cause premature fatigue failure of the pump shaft.

The pump efficiency at 50% of the BEP is considerably lower than at the BEP. Operating a pump in this manner will require more energy (higher operating costs), which could be offset by selecting a pump which may have a greater initial cost and operates at a lower speed and has a lower NPSH requirement.

This subject is covered in greater detail in ANSI/HI 9.6.3-1997 Centrifugal and Vertical Pump for Allowable Operating Region.


Q. I know that pumps should not be subjected to excessive forces and moments from the system piping, but how do excessive forces damage the pump and what is the weakest link in this regard?

A. Pump damage due to excessive forces and moments from system piping can cause damage in the following ways:

Shaft coupling misalignment â€“ Excessive pipe loads can distort the pump or its baseplate such that the alignment of the driver and pump shaft is forced beyond that which the coupling can tolerate without transmitting unacceptably high loads to the shafts. The result is shortened life of the coupling, bearings and mechanical seals.

Holddown bolt failure â€“ This can include elongation of the holddown bolts resulting in shaft movement at the coupling as well as reduced friction between the pump and baseplate causing the pump to move relative to its baseplate, thereby affecting the shaft coupling alignment.

Excessive stress in the pump nozzles â€“ The suction and discharge connections of the pump are already under stress due to the internal pressure in the pump case. Additional stress from the system piping can result in excessive stress in the casing structure. This is particularly dangerous with casings produced from a brittle material, such as cast iron.

Internal pump distortion â€“ End suction pumps, which are built with supports under the bearing housing, can be distorted by excessive piping

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strain which affects the internal alignment between the rotating and stationary parts. This may result in contact between these parts as well as misalignment of the mechanical seal faces. Premature seal failure and wearing ring damage may occur. Bearing life may also be affected.

Which of these problems becomes the weakest link depends on the specific pump design configuration and the direction or plane in which the forces and moments act. A recently published document, ANSI/HI 9.6.2â€“2001 Centrifugal and Vertical Pumps for Allowable Nozzle Loads contains recommended allowable loads for specific pump types and further explains the problem.


Q. We are designing a system with nine horizontal centrifugal pumps that will take suction from a common header. Is there a design standard or good practice that will provide information on the maximum water velocity in the suction header?

A. The Hydraulic Institute publishes ANSI/HI 9.8 â€“ 1998 Pump Intake Design, which provides information and design recommendations for suction piping.

â€œThe ideal flow entering the pump inlet should be of a uniform velocity distribution without rotation and stable over time.â€ �

â€œThe suction piping should be designed such that it is simple with gentle transitions in changing pipe sizes. Transitions resulting in flow deceleration at the pump shall not be used.â€ �

â€œThe maximum recommended velocity in the suction piping is 2.4m/sec or 8.0 ft/sec. Velocities may be increased at the pump suction flange by the use of a gradual reducer.â€ �


Q. Due to a recent thrust bearing failure in one of our deep well pumps, we would like to determine the down thrust on the bearing from the pump. Is there a simple way of calculating this?

A. Following is a simplified version of this calculation, which is found in Hydraulic Institute Standard, ANSI/HI 2.3 Vertical Pump for Design and Application. This equation is for impellers with no back wearing ring and applies to operation at BEP only.

The approximate down thrust at BEP in pounds is calculated as follows:

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Where H = Single stage head at BEP in feet s = Specific gravity of pumped liquid C = Experimental coefficient from figure 2.45 Adf= Area of impeller front wearing ring minus shaft area in in2

B = Number of bowls W = Total weight of all impellers and shaft in pounds

Figure 2.45 provides the experimental coefficient â€œCâ€, which includes the �impeller flow momentum change. This coefficient was obtained from a number of tests on vertical pumps with specific speeds from 1700 to 12000. The lines represent an average of these tests. A pump manufacturerâ€™s specific design may have a slightly different â€œCâ€ value.�


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Q. We are expanding our building, including the fire protection system. If the existing fire pump must be replaced, what special requirements must we look for in their replacement?

A. The purchase and installation of fire pumps must comply with strict regulations established by the National Fire Protection Association (NFPA) as published in their â€œPamphlet 20, Standard for the Installation of Stationary Pumps for Fire Protectionâ€ and â€œPamphlet 25, Standard for Inspection, �Testing and Maintenance of Water Based Fire Protection Systems.â€ �

More specific pump design requirements are also enforced by â€œUnderwriter Laboratoriesâ€ (UL) and â€œFactory Mutualâ€ (FM). These requirements � �include the following:

Pump type: Axially split case, end suction, vertical-in-line or vertical turbine pump.

Performance requirements: shut off head, head at 150% of rated flow, suction lift capability and maximum horsepower requirement.

Mechanical requirements: shaft strength, bearing life, no mechanical seals, materials of construction and hydrostatic test pressure.

Design approval by UL or FM based on their review of the pump design and witness of factory tests to insure that the pump performance meets their criteria. Approved pumps may then carry an FM or UL label when shipped.

A list of approved fire pump manufacturers can be found on the Hydraulic Institute web site www.pumps.org under Supplier Finder.


Q. With the ever increasing cost of oil and resulting increases in the cost of power, what can be done to decrease the operating cost of existing pump systems?

A. With existing pump systems, the most cost effective change that can be made is converting to variable speed drives for your pumps. At the same time, existing control valves must be removed. The result will be a reduction in the head loss in the system and lower operating speed for the pumps. See the figure on the right for an example of this effect. If you do the math, you should see a one or two year payback on the cost of conversion.

Other opportunities for power savings include: Maintain pumps close to new condition to avoid efficiency loss. Donâ€™t allow for excess margin in rate of flow or head. Review the pump selection and use a more efficient pump design if possible. Use two or more smaller pumps instead of one larger pump so that excess capability

can be turned off. Use pumps operating as turbines to recover pressure energy that would otherwise be

wasted.

The Hydraulic Institute publishes an education program on this subject titled, â€œEnergy Reduction in Pumps and Pumping Systems." Visit the e-Store at www.pumps.org for details.

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Q. I have heard that the rate of flow through centrifugal pumps can be regulated by submergence control instead of by throttling. How does this work, and does it cause damage to the pump?

A. Submergence control is sometimes used in applications where NPSH Available to the pump is limited and the pump would naturally be operating with marginal NPSH. A good example is condensate pumps in steam power systems.

Condensate pumps take condensed steam from the bottom of the steam condenser (hotwell). Ideally, when operating at the design conditions, the liquid level in the condenser hotwell provides sufficient NPSH to the pump so that it operates on the head curve, producing the design rate of flow. When the load on the electric generator is reduced, the steam required is reduced and the amount of condensed steam entering the condenser hotwell is reduced. This reduction in condensate causes the pump to draw down the level of liquid in the condenser, thereby reducing the NPSH available to the pump. The reduction of NPSH available to the pump causes it to cavitate, resulting in a reduction in the rate of flow. If the load on the generator does not change, the pump will operate in this mode until the load is increased back to normal.

The continuous cavitation in the impeller can be damaging to the pump, but this can be compensated for by selecting a more cavitation resistant impeller material, and using a more rugged design of the pump shaft and bearings.

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Q. I know that a centrifugal pump can overheat badly if run at shut off for some time, but how can I determine the minimum rate of flow through the pump to avoid excess temperature build up?

A. A commonly accepted practice limits temperature rise through a pump to 15Â°F. For most installations, this is adequate and minimum flow may be calculated with this equation:

Where:

Q = minimum flow rate, gpm;

Pp = input power at the minimum flow, hp;

2.95 = constant;

Cp = specific heat, BTU/lb-Â°F;

s = specific gravity.

At the minimum flows calculated using the above equation, the power input is approximately the same as at shut-off.

Catastrophic failure of the pump and associated equipment may result if the liquid within the pump casing is allowed to vaporize. To prevent flashing, a flow must be maintained through the pump which will keep the liquid below its saturation temperature.

Minimum flow is guaranteed by installing a bypass from the discharge line to some low-pressure point in the system. The bypass should not lead directly back to the pump suction.

An orifice installed in the bypass line breaks down the differential pressure between the pump discharge and the low-pressure point in the system.

The bypass may be manually or automatically operated but must be open during periods of light load or when starting or stopping the pump.

This subject is discussed in greater detail in ANSI/HI 1.3 Centrifugal Pumps for Design and Application.

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Pump FAQsÂ® November 2006 Article #3

Q. Mechanical seal failure represents most of the causes for pump repair in our plant, with unscheduled shutdown and loss of production. What can be done to relieve this problem?

A. Improved seal life can result from a review of failures with your seal supplier and making some changes to improve the seal designs and operation. Reducing unscheduled shutdown can be done by appropriate leakage detection to identify imminent failures before they occur. 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 detection is monitored to identify the failure mode of the seal or pressure boundary. These leaks may be in the form of liquid or vapor. Following are several means of monitoring leakage.

For less-hazardous liquids, leakage is often detected visually from joints or seal drains. Larger leaks of volatile light hydrocarbons such as propane may form ice deposits on the outside surface of the seal gland plate. Continued operation will cause the ice to melt and be replaced by carbon wear debris from the seal faces. Visual monitoring is commonly used for single and dual outboard double and tandem seals.

Sniffers are used to detect minute leakage of volatile organic compounds (VOCs). Typical locations monitored are joints, connections and seal drains. Concentrations can be measured to determine the severity of the leak. The proper sniffer must be used for the compound pumped. All single seal installations handling VOCs must use this method of monitoring.

Leakage through the inboard seal of a dual tandem 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 flare (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.

Leakage through the inboard seal of a dual tandem seal arrangement may be detected by monitoring the gas flow from the seal to the flare system. Leakage through the inboard seal of a dual double seal arrangement may be detected by measuring the loss of barrier liquid from the circulation system and reservoir. The consumption of barrier gas through a dual double gas-lubricated seal will vary with changes to pressure, temperature and speed.

For more detail on this subject, see HI Standard ANSI/HI 9.6.5 Centrifugal and Vertical Pumps for Condition Monitoring

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Pump FAQsÂ® November 2004 Article #2

Q.

In our power plants we have several applications where we need to reduce the NPSH Required by the pumps by as much as fifty percent. Is this possible? If so, how can this be done?

A. The NPSH required by a pump is a function of the pump operating speed and rate of flow required. When these criteria are set in a given design, little can be done to reduce NPSH Required. The only thing that can make a significant is the addition of an inducer. Inducers are devices designed to benefit the functioning of the impeller by increasing the liquid pressure before it enters the impeller. See figure 1.59, which is from ANSI/HI 1.3-2000 Centrifugal Pumps for Design and Application. Even this device does not change the basic function or NPSH Required by the impeller, it simple increases the available inlet pressure.

When properly designed and matched to the impeller, the NPSH Required by the pump with the inducer can be reduced by as much as 50%. However, most manufacturers do not make them available.

Regarding further reading on this subject, there is so much available that it is difficult to identity a starting point. A literature search service or an internet search should be able to help.

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Q. We have a new water system pump which operates with the excessive noise with smaller existing pumps. The pump produces 7300 gpm and is driven by a 900-hp motor. The system provides 23.5 feet of NPSH and only 18 feet is required by the pump. The noise sounds like the pump is handling sand. What is the cause of the noise, and can it be eliminated? After nine months of operation, examination shows no impeller damage.

A. A pump is a hydraulic machine, and all mechanical devices are inherently noisy to some extent. Recent studies by Hydraulic Institute members show a

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strong correlation between noise and pump power. Higher horsepower pumps are naturally more noisy. If there was no apparent impeller damage after nine months of operation, the pump can be expected to operate indefinitely.


Q. We recently rebuilt a vertical turbine pump with new impellers and wearing rings and tested its performance. The pump is handling water at ambient temperature (S=1.0). However, the result showed total head performance which was lower than the published curve as the rate of flow was increased. The pump is thirty feet long so we put a test gage two pipe diameters downstream of the discharge elbow. The shut off head nearly matches the published curve. Do you have any suggestion for increasing the pump head?

A. The performance curve for vertical turbine pumps is usually for the bowl assembly only and does not include the head losses in the pump column pipe and discharge head. In order to measure the bowl performance, the discharge pressure gage must be connected to the pump column pipe two diameters above the outlet from the bowl assembly. However, the gage can be located above ground and connected to the lower column pipe with a small diameter (1/4 inch ) tube. See figure.

Note that the height of the gauge above the first stage impeller must be measured as well as the inside diameter of the column pipe. The total head with water can then be calculated as follows:

Hba = 2.31pgba + Zd â€“ Zw + vd 2/2g where:

Hba = bowl assembly head â€“ feet of water

pgba = discharge gage pressure-psi

Zd = height of discharge pressure gage above first stage impeller datum â€“ feet

Zw = height of water level above first stage impeller datumâ€“ feet

vd = liquid velocity in column pipe â€“ feet/second

An alterative to this approach is to calculate the losses in the column pie and discharge elbow using published data on flow friction losses, such as the Hydraulic Institute Engineering Data Book or computer software programs and adding such values to the measured pump head.

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Q. The FAQ in the February 2004 issue of P&S showed the performance curve of a centrifugal pump when handling entrained air. Do self priming pumps perform the same way when pumping liquid with entrained air, or can they do better?

A. During the priming cycle, self-priming pumps are handling 100% air which is at much higher percent than a centrifugal pump could handle, but not at the same head as when pumping water. When priming, self-priming pumps must be vented to the atmosphere to allow the air to be expelled . This is usually accomplished with the discharge pipe empty or with an air release valve on the pump.

During the pumping cycle, the self priming pump will suffer a similar reduction in total head as a regular pump but it should be able to handle a greater percentage of air before pumping stops.


Q. Most industrial pump manufacturers publish centrifugal pump performance curves that include a curve for NPSHR. I understand that this curve is based on tests, which determine the NPSH value when the total head is reduced by 3%. Isnâ€™t it better to publish NPSHR curves with higher values based on 0%

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reduction in total head or zero cavitation?

A. As NPSH available to a centrifugal pump increases, the severity of the cavitation forces actually increases until NPSHA is about 2 times NPSHR. In order to reach a level of zero cavitation, values as great as 10 to 20 times the NPSHR are required. In many applications such high values are impractical to design for. The severity of cavitation damage to a pump also depends on the properties of the liquid pumped and the material of the impeller. For example, hydrocarbon liquids usually cause no damage when the pump operates with NPSHA equal to NPSHR. Regarding materials, stainless steel and aluminum bronze have been found to be more resistant to cavitation damage. Increasing the values of NPSHR would eliminate the opportunity to take full advantage of these variables.

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Pump FAQ’s by Hydraulic Institute (UK)

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