Pump Basics

Pump - Basics

Lesson Four - The Affinity LawsThe affinity laws of centrifugal pump performance express the effect on pump performance due to changes in certain application variables. The affinity law variables which affect pump performance are:

1) Pump speed in pump revolutions per minute (RPM).

2) Impeller diameter.

Changing the Pump Speed (RPM):

When the impeller diameter of a centrifugal pump is held constant the effect of changing the speed (RPM) of the pump is in accordance with the following:

Capacity: Q1/Q2 = N1/N2

Head: H1/H2 = (N1/N2)2

BHP: BHP1/BHP2 = (N1/N2)3

Where subset number 1 shows performance at the initial speed and subset number 2 shows performance at the new speed, and:

Q = Capacity, GPM

H = Head, Feet

BHP = Brake Horsepower

N = Pump Speed,RPM

Analysis: As shown from the above equations, changing the speed (RPM) of a pump affects the flow, head and input brake horsepower of the pump in different proportions. Changing the speed affects the flow through the pump by a proportion equal to the increase or decrease in speed. The pump head is changed by the square of the proportion of speed change, while the brake horsepower is changed by the cube of the proportion of speed change.

Changing the Impeller Diameter:

When the speed (RPM) of a centrifugal pump is held constant the effect of changing the impeller diameter (D) is as follows:

Capacity: Q1/Q2 = D1/D2

Head: H1/H2 = (D1/D2)2

BHP: BHP1/BHP2 = (D1/D2)3

Where subset number 1 shows performance at the initial impeller diameter and subset number 2 shows performance at the new impeller diameter, and:

Q = Capacity, GPM

H = Head, Feet

BHP = Brake Horsepower

D= Impeller Diameter

Analysis:

Analysis: As shown from the above equations, changing the diameter of a pump impeller affects the flow, head and input brake horsepower of the pump in different proportions. Changing the impeller diameter affects the flow through the pump by a proportion equal to the increase or decrease in diameter. The pump head is changed by the square of the proportion of diameter change, while the brake horsepower is changed by the cube of the proportion of diameter change.

Why NPSHR Changes With Impeller Diameter?Introduction

Confusion sometimes results when reviewing published NPSHR curves.  This is especially true when faced with trimming the impeller diameter to match changing operating conditions.  A well known fact is that the head-flow relationship varies with the diameter.  This can be accurately approximated by the affinity laws.  However, what happens to the NPSHR-flow relationship when the diameter changes?  This relationship is frequently over looked and can lead to pump cavitation.  This Tip of the Month examines the relationship of NPSHR to the impeller diameter and clarifies other misconceptions regarding pump NPSHR curves.

Background

Pump performance may be shown for a single impeller or a range of impeller diameters.   In the latter case the pump performance may be shown as multiple curves from the maximum to the minimum diameter, and may show several intermediate impeller sizes.  In addition, the pump performance characteristics may show curves for NPSHR, efficiency and required power.   The representation of the pump performance varies widely depending on many factors and can lead to design errors and possible confusion.

An example of a typical pump performance curve frequently seen in publications is shown in Figure 1.  The pump flow rate is plotted on the horizontal axis, and the head and NPSHR curves, which are a function of flow rate, plotted on the vertical axles.  Note that a single-line NPSHR curve starts at the no-flow condition and continually rises to the maximum flow rate.  For several reasons that will be discussed later, this type of NPSHR curve is incorrect and can lead to design errors and possible cavitation problems.

Pump cavitation is a complex subject and the topic of many technical papers and books.  However, it is widely accepted that this phenomenon begins at the pump inlet.  It basically results from the increased velocity and reduced pressure as the fluid enters the impeller.  If the fluid static pressure drops below the vapor pressure, gas bubbles form and later collapse as the fluid flows along the impeller vanes.  These vapor bubbles can have a significant effect on the head produced by the pump.

It is important to note that fluid temperature also plays an important part in pump cavitation.   Obviously, the fluid vapor pressure will vary with temperature.  The fluid temperature will also vary with pump efficiency.  Temperature rise due to pump efficiency is not significant in the high to mid-range flow rates, however, can be very significant at low flow rates.   This is why pump NPSHR values are not given at low flow conditions.

Figure 1 – Pump Performance Curve for a Range of Impeller Diameters

Another important factor in pump cavitation is the fluid velocity.  Fluid entering a pump will continually increase in velocity as it passes to the impeller eye.  This increase in velocity causes a drop in the fluid static pressure and is analogous to lift on an airfoil.    At high to mid-range flow rates the incoming fluid velocity and the impeller rotational velocity are compatible and contributes to stable flow through the pump.  However at low flow rates the entering velocity is well below the rotational velocity and may cause the fluid to “recirculation” at the impeller inlet.  Fluid recirculation is another form of pump cavitation. This is another reason why NPSHR is not given at low flow rates.

NPSHR Testing

Understanding how NPSHR tests are conducted and how the impeller diameter influences the produced head will help eliminate confusion and possible errors.  Pump manufacturers determine the characteristic shape of the NPSHR curve for each impeller through carefully controlled shop testing, hydraulic modeling and computer simulation.  Hydraulic Institute Standard 1.6 gives strict guidelines for conducting shop testing and is used by most pump manufactures.  Pumps are normally connected to closed-loop piping circuit where water flows from a suction tank (or sump) through the pump and then back to the tank. The discharge flow rate, temperature and pressure are carefully measured and controlled throughout the test.    Basically the test is conducted at a fixed flow rate and speed while the suction pressure is reduced.   By reducing the suction pressure a point is reached when the water begins to vaporize thus causing the pump to cavitate.  The characteristic “cavitation” point is the flow rate that is exhibited by a small drop in head.  The test is conducted again at another fixed flow rate and again the resulting suction pressure and flow rate value are recorded at the “cavitation” point.  Once the series of tests are completed, a smooth line is drawn through the recorded data and plotted.   Figure 2 illustrates a typical series of test results and the resulting NPSHR curve.

Figure 2 – NPSHR Test Curve

A pump cavitation point can be difficult to define.  The formation of vapor bubbles is a gradual process, starting slowly and increasing with flow rate.  The API-610 defines the cavitation point as a three percent drop in head.  This is not to say that pump cavitation does not occur at smaller values, it is just difficult to accurately measure at smaller values.     To obtain a single point it is necessary to run a pump for a period of time and allow the testing circuit to stabilize to the reducing suction pressure.  Remember, vapor bubbles are forming and instruments need time to react to the fluid dynamics.

Impeller Diameter and Head Relationship

Larger pump impellers produce greater values of head for a given speed.   This is because the head is proportional to the tip speed.  The relationship of head to tip speed can be approximated by Equation 1.

(Eq. 1)

Tip velocity can also be related to impeller diameter and rotating speed by Equation 2.

(Eq. 2)

From Equations 1 and 2 it can be seen that changes in impeller diameter will have a direct effect on the pump head.  For example, reducing the impeller diameter will lower the pump head by a factor of four.   Since the cavitation point is identified by a three percent drop in pump head, it is logical that any change in impeller diameter will have a direct effect on the NPSHR value.  For this reason, most pump manufacturers provide a single NPSHR curve for a given impeller diameter.  Figures 3 and 4 are typical pump performance curves for a range of impeller diameters.  Note that a separate NPSHR curve is given for each diameter.

Figure 3 – Typical Pump Performance Curve for a Range of Diameters

Figure 4 – Optional Pump Performance Curve for a Range of Diameters

Conclusions

The following conclusions can be reached from the previous discussion.

1. Each impeller will have a characteristic NPSHR curve. It will depend on many design factors including the diameter.

2. At a given flow rate, the NPSHR increases as the impeller diameter is reduced.

3. The NPSHR is never tested at the shut-off point. The fluid temperature continually rises as the flow rates decreases. This prevents the system from stabilizing sufficiently to obtain accurate measurements.

4. Pumps may cavitate at low flow rates due to recirculation of fluid at the impeller eye.

5. The shape of the NPSHR curve is a U-shape. There is a slight rise in values as the flow is reduced and again at higher values. The NPSHR is lowest in the mid-range values.

By: Joe Honeywell

Legend

A          Conversion constant = 720 ft/sec (600 m/s)

D          Impeller diameter, inches (cm)

H          Total pump head, ft (m)

g          Gravitational constant, 32.17 ft/sec2 (9.81 m/s2)

n          Rotational speed, rev/min

V          Impeller tip velocity, ft/sec (m/s)

References

1. American Petroleum Institute Standard 610, Centrifugal Pumps for Petroleum, Petrochemical and Natural Gas Industries, 10th Ed.

2. Hydraulic Institute Standard 1.6, Centrifugal Pump Tests, 2000

3. Terry Henshaw, Pumps and Systems, May 2009

Written by Joe HoneywellMr. Joe Honeywell is a graduate of University of Tulsa with a Bachelor of Science in Aerospace Engineering and a Master’s of Science in Mechanical Engineering. Mr. Honeywell began his career with an engineering consulting company named Crest Engineering. He worked in the mechanical department for thirteen years, specializing in rotating equipment, pressure vessels and piping systems. Mr. Honeywell advanced to project engineer and later project manager, where he was involved in many offshore and onshore projects for oil and gas producers, both domestic and international. Mr. Honeywell joined another consulting company, Crown Tech, Inc. where he worked for 19 years and became a principle in the company. At CTI, Mr. Honeywell provided engineering services to many oil and gas producers, pipeline companies, power producers and equipment manufacturers. His responsibilities included project management, design, manufacturer, construction management, start-up and operation of power plants, pipelines and production facilities. Mr. Honeywell’s background includes extensive experience with mechanical systems and rotating machinery. Mr. Honeywell is a Registered Professional Engineer; and a member of ASME and holds a U.S.A. patent.

Impeller NPSHR

Q  -  Does the NPSHR of an impeller change when its diameter is cut?

A  -  No. The impeller NPSHR is dependent on its suction geometry such as the eye diameter, eye area, and vane inlet angle. Cutting the impeller diameter has no effect on the suction geometry, and should have no effect on NPSHR.

Q  -  Most pump performance curves shows the same NPSHR curve regardless of impeller diameter but there are also some that show two NPSHR curves: one for maximum impeller diameter and another for minimum diameter. I noticed that the NPSHR for minimum diameter is higher than for maximum diameter. You said that NPSHR does not change with impeller cut diameter. Please explain this apparent inconsistency.

A  -  You raised a very interesting point. Our sources think that this has more to do

with how NPSHR is determined rather than with actual change in NPSHR. These explanations were given:

This is an anomaly that is caused by internal flow recirculation and how flow rate is measured. In some instances cutting the impeller diameter will result in higher internal flow recirculation. Radial flow impellers whose ratio of eye diameter to impeller diameter is greater than 0.50 are more sensitive to this phenomenon.

For clarity, assume a hypothetical situation where a pump has:

Qd, flow at discharge nozzle = 100 GPMQi, internal flow recirculation = 10 GPMQs, flow at impeller suction, Qd+Qi = 110 GPMDuring test the NPSHR is measured at 100 GPM discharge flow rate. That NPSHR is considered the NPSHR at 100 GPM but in reality it applies to 110 GPM, the actual flow rate at the impeller suction, of which 10 GPM is due to internal recirculation.

Now assume that the impeller diameter is cut and, as a result, the internal recirculation flow increases to 15 GPM. The NPSHR at 100 GPM discharge flow will now appear to be higher because in reality it is the NPSHR for 115 GPM suction flow.

The problem is that there is no practical way to measure internal flow recirculation and hence there is no practical way to correct the NPSHR.

Another explanation:

It is common practice in the pump industry to measure NPSHR based on a 3% head loss. But 3% of what head? If a pump is tested at maximum impeller diameter then it is 3% of the head at maximum impeller diameter; if tested at minimum diameter, then it becomes 3% of the head at minimum diameter. In my opinion this practice is inconsistent and needs to be corrected.

Here's why: Say that a pump, at 1000 GPM, has a head of 500 FT at maximum impeller diameter. Under the present industry-wide practice a head drop of 15 FT will determine its NPSHR. But if that same impeller is cut 20% of its diameter and the head at 1000 GPM is reduced to 300 FT then its NPSHR will be based on a head drop of only 9 FT.

See the inconsistency? The same pump, with same impeller and same suction geometry, ends up with different NPSHR at same capacity because the NPSHR are based on different absolute values of head drop. This seems to explain why the NPSHR at impeller cut diameter may at times appear to be higher.

Specific speed (NS) and suction specific speed (NSS) are always calculated based on data at maximum impeller diameter for data consistency. It makes sense that NPSHR should also be based on 3% head drop at maximum impeller diameter regardless of actual impeller cut diameter.

R: 0210-IMNPC: basics, operationF: NPSHRSUBJECT: All about impellers 10-1

First the types:

The open impeller is nothing more than a series of vanes attached to a central hub for mounting on the shaft without any form of side wall or shroud. This design is much more sensitive to vane wear than the semi or closed impeller.

The semi-open impeller incorporates a single shroud at the back of the impeller. This is the most common design used in the United States and the one you find on most ANSI standard pumps.

o The shroud often has "cast in" pump out vanes that will help circulate lubricating liquid from the lantern ring connection through the packing ahead of the lantern ring.

o Most modern pump designs allow you to adjust the semi- open impeller without disassembling the pump. This is a tremendous advantage if you want to maintain the pump efficiency by adjusting the impeller to volute clearance for thermal expansion and volute/impeller wear. Remember that if there is a mechanical seal in the stuffing box any impeller adjustment can interfere with the seal face loading. Those designs that adjust to the volute (Goulds type) will unload the seal faces and those that adjust to the back plate (Flowserve or Duriron type) will increase the seal face loading.

o A typical volute or back plate clearance for a semi open impeller would be 0.015 to 0.020 inches (0,4 to 0,5 mm). For each 0.002 inches (0,05 mm) you increase this clearance, the pump will lose about 1% of its capacity.

The closed impeller has a shroud on either side of the vanes. This is the most common design found with ISO standard pumps, oil refinery applications and the design you see on double ended pumps.

o To maintain impeller efficiency you are required to replace the wear rings after the original clearance has doubled. The first problem is to determine when it has doubled, and then you have to take the pump apart to replace them. The result is that timely replacement is seldom done, and pump loss of efficiency with resultant vibration becomes the rule.

The general rule of thumb is that the pump will lose about 1% of its capacity for each excessive 0.001 inches (0,025 mm) of impeller clearance.

Since the wear ring clearance is usually smaller than the area of the balance holes drilled through the impeller, you will lose the advantage of suction recirculation as stuffing box pressure is very close to suction pressure.

The impeller specific speed number describes the shape of the impeller

The shape of the head/ capacity curve is a function of specific speed, but the designer has some control of the head and capacity through the selection of the vane angle and the number of vanes.

The pump with the highest specific speed impeller, that will meet the requirements of the system, probably will be the smallest and the least expensive. The bad news is that it will run at the highest speed and be subject to maximum wear and damage from cavitation.

Radial flow impellers (low specific speed numbers)

They should be specified for high head and low flow conditions. They seldom exceed 6 inches (150 mm) in diameter and run at the higher motor

speeds

The casing is normally concentric with the impeller as opposed to the volute type casings normally found in the industry..

These impellers exhibit a flat head/capacity curve from shut off to about 75% of their best efficiency and then the curve falls off sharply.

Radial flow impellers are normally started with a discharge valve shut to save start up power.

Axial flow impellers (high specific speed numbers)

They run at the highest efficiency They have the lowest NPSH requirement.

They require the highest power requirement at shut off, so they are normally started with the discharge valve open.

Impellers can be manufactured from a variety of materials:

We would like a combination of a hard material to resist wear and a corrosion resistant material to insure long life. This is often a conflict in terms because when we heat treat a metal to get the hardness we need, we lose corrosion resistance. The softer metals can have corrosion resistance, but they lack the hardness we need for long wear life. The best materials that combine these features are called the "Duplex Metals". These duplex materials are now in their second generation. They can be identified by letters and numbers such as Cd4MCu

If a new impeller is required because of cavitation, the new design should incorporate those features we have learned that will increase impeller performance:

The use of large fillets where the vanes join the shrouds to lessen stress. Investment castings so that you can design in the compound curves that produce less

wear.

The latest design iteration to help reduce radial thrust.

Sharpened leading edges of the vanes to reduce losses.

A reduction of shroud to cutwater clearance to lessen internal recirculation.

A conversion to the newer duplex metals.

Impellers can be designed for a variety of applications:

The ideal impeller would have an infinite number of vanes of an infinitesimal size. The conventional impeller design with sharp vane edges and restricted areas is not

suitable for handling liquids that contain rags, stringy materials and solids like sewage because it will clog. Special non-clogging impellers with blunt edges and large water ways have been developed for these services.

Paper pulp impellers are fully open and non-clogging. The screw conveyer end projects far into the suction nozzle permitting the pump to handle high consistency paper pulp stock.

Vortex pump designs have recessed impellers that pump the solids by creating a vortex (whirl pool effect) in the volute and the solids move without ever coming into contact with the impeller. You pay for this feature with a greater loss of pump efficiency.

An axial flow impeller called an Inducer (it works like a booster pump) can be placed ahead of the regular pump impeller, on the same shaft, to increase the suction pressure and lessen the chance of cavitation. In some instances this can allow the pump to operate at a higher speed with a given NPSH. The inducer will contribute less than 5% of the total pump head, and although low in efficiency the total efficiency of the pump is not reduced significantly. The total reduction in NPSH required can be as much as 50%.

People often inquire about forward curved vanes. Tests have shown:

Both the capacity and efficiency were reduced. There was a slight increase in head.

The impeller exhibited unstable characteristics at the low end of capacity range.

The impeller exhibited steep characteristics at high end of the range.

Increasing the number of vanes tends to flatten out the curve and steady the flow.

Impellers can be single or double suction designs.

Because an over hung, single suction impeller does not require an extension of the shaft into the impeller eye it is preferred for applications handling solids like sewage. The suction eye is defined as the inlet of the impeller just before the section where the vanes start. In a closed impeller pump the suction eye is taken as the smallest inside diameter of the shroud. Be sure to deduct the impeller shaft hub to determine the area.

Double suction pumps lower the NPSH required by about forty percent.

o Most double suction impellers are constructed so that the stuffing box is at suction pressure. This causes you to lose the advantage of suction recirculation to prevent seal failure when handling solids. You are going to have to flush

many of these seals with a clean, compatible liquid that will dilute your product to some degree.

Looking at the axial thrust in single stage pumps.

o The axial thrust generated is higher than in closed impellers because of the hub. Pump out vanes and balance holes are a common solution to this problem.

o A mechanical seal can add to this axial thrust. The amount is dependent upon the design of the seal. Balanced designs create less thrust.

o Balancing holes are not desirable with closed impellers because leakage back to the impeller inlet opposes the main flow creating disturbances. A piped connection to the pump suction can replace the balance holes

o Theoretically there shouldn't be any thrust in a double suction closed impeller, but:

An elbow with the inlet piping running parallel to the shaft will cause an uneven flow into the impeller eyes. This uneven flow will cause thrusting of the impeller in one direction depending upon the flow difference. The eye is taken as the smallest inside diameter of the shroud. Remember to deduct the area occupied by the impeller hub.

The two sides of the discharge casing may not be symmetrical causing an axial thrust.

Unequal leakage through both sets of packing can upset the axial balance. Leaking seals can do the same thing.

Impellers can be cut down to keep the application close to the pumps best efficiency point :

Theoretically up to twenty five percent of an impeller diameter can be removed, but any time you remove more than ten percent of the maximum impeller diameter the affinity laws are no longer accurate because of slippage between the impeller outside diameter and the pump volute.

Changing the impeller diameter changes the head, capacity and power requirements.

The capacity can be increased by under filing the vane tips, but the discharge head and the power requirement will automatically adjust to the values where the pump curve intersects the system curve.

If you intend to cut down the impeller diameter, the impeller should be cut down in at least two steps and tested after each step.

After cutting down the impeller diameter the discharge vanes should be reshaped to a long gradual taper to increase the pumps performance. Chamfering or rounding the discharge tips will frequently increase the losses and should never be done.

Over filing is removing metal from the leading edge of the blade. This seldom produces any increase in the vane spacing and produces a negligible change in pump performance.

Under filing is removing metal from the trailing edge of the blade. If properly done it will increase the vane spacing and can increase the capacity by as much as ten percent.

If the inlet vane tips are blunt, over filing will increase the inlet area and the cavitation characteristics can be improved.

Cutting back the tongue increases the throat area and increases the maximum capacity. The head/capacity is then said to "carry out further".

Q. What is NPSH?     A. More often the term NPSH is misinterpreted as suction lift or suction head. This is no

correct. The definition of NPSH is the absolute energy (Head) required to push the liquid into the impeller eye. If this is not provided, then the pressure at the eye of the impeller drops below the vapour pressure of the liquid and cavitation takes place. This ultimately deteriorates the performance of the Pump. The available NPSH can be worked out by the use of following formulae. See that the available NPSH is more than the NPSH required. (Specified by the Pump Manufacturer)

NPSHA = Ha ± Hst - Hvp - Hf.

    Where     Ha = Atmospheric Pressure expressed in terms of liquid column.HST = Negative / Positive Static Head.Hvp = Vapour Pressure of the liquid expressed in terms of liquid column.Hf = Frictional Head losses in suction pipeline.        Q. What is the difference between NPSH and Suction Lift?    

A. Suction Lift is a negative head (liquid level below Pump Center Line), the Pump has to overcome. Whereas the NPSH is net positive suction head available to the Pump at Impeller eye. If available NPSH is more than required NPSH (Specified by the Pump Manufacturer), then Pump will work satisfactorily.

    Q. What is the effect of density of the liquid on the performance of the Pump?   A. Pressure developed and power consumed by the Pump is directly proportional to the

density of the liquid. This is explained as under.If a Pump develops 1kg / cm2 pressure and consumes 1 Kw power for a density of 1 g / cm3 the pressure developed by the Pump for another liquid of a density of 1.5g / cm2 and power consumed will be 1.5 Kw.

    Q. What is the effect of viscosity of the liquid on performance of the Pump?   A. Higher viscosity of the liquid will have an effect on parameters like Head, Rate of flow

Efficiency. The conversion factors can be derived for theses parameters from the graphs provided by Hydraulic Institute.

        Q. What is the effect of altitude on the performance of the Pump?     A. Atmospheric pressure is dependent upon the altitude. Higher the altitude less than

atmospheric pressure. Therefore available NPSH will get reduced at higher altitudes. This has to compensate y providing higher static head on suction side of the Pump.

      Q. What is the effect of vapour pressure on the performance of the Pump?   A. Vapour pressure value will have direct effect on the available NPSH. Higher the vapour

pressures lesser the NPSH available. This has to be compensated by providing higher static head on suction side of the Pump.

        Q. What is the effect of speed on the performance of the Pump?    

A. All the performance parameters such as Head, Rate of flow and Power of the Pump will change due to change in speed. The changed parameters can be worked out by using following formula

The Rate of Flow will change in direct proportion to the ratio of speed.

Q1 / Q2 = N1 / N2

The Head will change in square proportion to the ratio of speed.

H1 / H2 = (N1 / N2)2

The Power will change in cube proportion to the ratio of speed.

P1 / P2 = (N1 / N2)3

    Where

Q1, H1, P1 are known parameters at known speed N1.Q2, H2, P2 are parameters to be calculated at desired speed N2

        Q. What is the effect of impeller diameter on the performance of the Pump?     A. All the performance parameters such as Head, Rate of flow and Power of the Pump will

change due to change in impeller diameter. The changed parameters can be worked out by using following formula.

The Rate of Flow will change in direct proportion to the ratio of impeller diameters.

Q1 / Q2 = d1 / d2

The Head will change in square proportion to the ratio of impeller diameter

H1 / H2 = (d1 / d2)2

The Power will change in cube proportion to the ratio of impeller diameters.

P1 / P2 = (d1 / d2)3

    Where Q1, H1, P1 are known parameters at known impeller diameter d1.Q2, H2, P2 are parameters to be calculated at desired impeller diameter d2.

        N.B.: These conversion formulas generally hold good for Pumps having Radial Flow Impellers. Do not use these formulas for mixed flow and action flow impellers.       Q. What is the effect of direction of rotation on the performance of the Pump?    

A. Many people think that Pump will not work if it is run in reverse however it is not true. If the Pump is run in reverse rotation it will develop Approx. 80% rate of flow at 65% head and will run at 50% of the rated efficiency. Thus the Performance of the pump deteriorates substantially and hence not advisable to run the Pump in reverse direction.

        Q. What is the effect of higher size motor rating of same speed, on the performance of the Pump?    

A. Higher size motor rating will not change the performance of the Pump as the speed of the motor is same. Hence it is not advisable to use higher size motor rating of the same speed to improve the performance of the Pump.

HELP FOR CALCULATING THE NEW IMPELLER DIAMETER OR SPEED FOR A NEW FLOW RATE

Applets are programs based on the java language that are designed to run on your computer using the Java Run Time environment.

This program is designed to help you calculate the diameter of the impeller required to achieve a permanent change in your system's flow rate. The calculations are based on the affinity laws which in turn are derived from a dimensionless analysis of three important parameters that describe pump performance (ref: The Pump Handbook by McGraw-Hill, chapter 2). The analysis is based on the reduced impeller being geometrically similar and operated at dynamically similar conditions or equal specific speed.

The affinity laws or equations are:

where subscripts 1 and 2 denote the value before and after the change. P is the power, n the speed, D the impeller diameter , H the total head.

If the speed is fixed they become:

If the diameter is fixed they become:

The process of arriving at the affinity laws assumes that the two operating points that are being compared are at the same efficiency. The relationship between two operating points, say 1 and 2, depends on the shape of the system curve (see next Figure). The points that lie on system curve A will all be approximately at the same efficiency. Whereas the points that

lie on system curve B are not. The affinity laws do not apply to points that belong to system curve B. System curve B describes a system with a relatively high static head vs. system curve A which has a low static head.

Diameter reduction To reduce costs pump casings are designed to accommodate several different impellers. Also, a variety of operating requirements can be met by changing the outside diameter of a given radial impeller. Euler's equation shows that the head should be proportional to (nD)2 provided that the exit velocity triangles remain the same before and after cutting. This is the usual assumption and leads to:

Which apply only to a given impeller with altered D and constant efficiency but not a geometrically similar series of impellers.If that is the case then the affinity laws can be used to predict the performance of the pump at different diameters for the same speed or different speed for the same diameter. Since in practice impellers of different diameters are not geometrically identical, the author's of the section called Performance Parameters in the Pump Handbook recommend to limit the use of this technique to a change of impeller diameter no greater than 10 to 20%. In order to avoid over cutting the impeller, it is recommended that the trimming be done in steps with careful measurement of the results. At

each step compare your predicted performance with the measured one and adjust as necessary.

The applet is designed to be used in English or French.

Hydraulic re-rate

All too often an existing pump can become too big, too small, or unfit for its current service because of changes in its operating conditions. This results in the machine becoming a “bad actor pump” – to use a commonly used phrase in the industry.

One solution to this “bad” situation is to consider replacing the equipment with a new unit that is right-sized for the current service. Another solution is to consider a hydraulic re-rate in lieu of buying a new unit.

Hydraulic re-rate refers to changing the rated operating conditions of a pump by modifying its hydraulic parts, or components, to effect the change.

The rated conditions can be the machine’s operating capacity, differential head, and speed; the liquid's specific gravity, viscosity, temperature, suction pressure, and vapor pressure; or the site's net positive suction head available (NPSHA).

Hydraulic parts are parts that are in contact and in the main path of the pumped liquid as it enters the suction nozzle and exits the discharge nozzle.

They include the impeller, volute (or diffuser), the suction case or suction bay of the casing, the cover if it were part of the suction bay, and the suction and discharge nozzles. The design drawings depicting these parts are aptly referred to as hydraulic drawings.

In multistage pump, it may include the short crossover, intermediate crossover (if the first stage impeller were of double suction design), the long crossover, and the final discharge section. In a barrel or double case pump it may include a separate suction case, and the barrel.

From hydraulic re-rate standpoint wear rings, shaft sleeves, and balance disks are not normally considered hydraulic parts because their design has no bearing in determining or setting the rated conditions although these parts may have to be changed also during a re-rate for axial thrust balancing. The design drawings depicting these parts are referred to as machining drawings – they are not hydraulic drawings

Advantages of hydraulic re-rate

It results in higher hydraulic efficiency and in reduced energy consumption. It improves the reliability and prolongs the useful life of the equipment.

The re-rate can be done faster and can be scheduled to minimize disruption to plant operations.

It minimizes risks and hazards, and improves safety to equipment and personnel.

The decision process to do a re-rate is simpler and faster because it may not require the review and approval of others in the corporate hierarchy.

A re-rate typically does not require expensive and time-consuming modifications to existing base and piping that may be required of a new equipment.

Existing spare parts on stock may still be usable.

Depending on location or legal jurisdiction, re-rated equipment may not be subject to the same stringent environmental compliance, such as emission limits, that typically apply to new equipment.

Operations and maintenance personnel are already familiar with the equipment – no additional training, or learning curve, is needed.

Disadvantages of hydraulic re-rate

The modification required to do the re-rate has to be done within the physical confines of the existing parts or components. The pump may not achieve optimum performance or efficiency for the service compared to new equipment. Compromises in the design may have to be made because of existing physical or dimensional constraints.

Its life cycle, although may be extended due to improved reliability, is expectedly not as long as that of new equipment.

Changes or modifications

Some of the changes or modifications that are typically made to perform a hydraulic re-rate are:

Change the impeller with one of lower or higher flow design. Change the impeller with one of lower or higher differential head design.

Change the impeller with one of different NPSHR characteristic due to a change in site NPSHA.

Modify the volute to decrease its throat area to hydraulically match a decrease in rated flow rate.

Modify the volute to increase its throat area to hydraulically match an increase in rated flow rate.

Increase (or reduce) the impeller diameter to increase (or reduce) its differential head using the Affinity Laws to estimate the new diameter

Increase (or reduce) the rotor speed to increase (or decrease) its differential head using the Affinity Laws to estimate the new speed.

De-stage, or up-stage, a multistage pump to change its differential head. De-staging refers to the removal of one or more impellers; up-staging is the reverse of de-staging.

Not all centrifugal pumps are good candidates for hydraulic re-rates. The costs of replacing small ISO, ANSI, or ASME pumps with new units are probably less expensive than the combined cost of engineering and replacement parts needed to do a hydraulic re-rate. In general, most API centrifugal pumps, and bigger sizes of ISO, ANSI, and ASME pumps are good candidates.

Questions:

1. What are the characteristics of a high flow impeller, relative to an existing impeller, if one were to perform a high flow hydraulic re-rate? What about a low flow impeller?

2. What is the practical limit in flow rate by which the BEP of a given impeller can be shifted to the right of its performance curve? What about to shifting the BEP to the left?

3. What precautions should be taken when chipping, or cutting, back the volute lips to increase its volute throat area?

4. Tests have shown that a modification to an impeller casting can be made to increase its head rise to shut off. What is that modification?

5. What is the main advantage of a double suction impeller with staggered vanes over a typical double suction impeller with non-staggered vanes? What is its main disadvantage?

6. What precautions should be taken when V-notching an impeller? How does V-notching the impeller outlet vanes help alleviate the vibration at impeller vane pass frequency?

7. What precautions should be taken when chipping or cutting back volute lips?