122574668-17027756-Pump-Learning-Guide1

-The problem with pump standards

All about specific speed How to read a pump curve?

Why prime a centrifugal pump? The oversized pump Understanding the system curve

How efficient is your pump? Piping system for pumps

Calculating the total system head

What do we mean by pump efficiency?

Prevent potential seal and pump problems

Estimating the shutoff head for pumps

Rules of thumb for pumps Rules of thumb for seals

Calculate the water HP coming out of pump

A new technique of troubleshooting

The best pump and sealtechnology?

Operational practices to avoid problems

Pumps does not develop enough head?

Pump does not give enoughflow?

Pump selection practices leads to problems

OSHA 1910 regulation

The problem with pump standards:

A Quick check of existing pump standards will reveal that there are a variety of them. The list includes:

• Hydraulic Institute Standards • American National Institute Standards for Chemical Pumps :

o B73.1 for Horizontal type. o B73.2 for Vertical Inline

• API 610 for centrifugal Pumps • API 674 for Reciprocating Pumps • API 675 for Controlled Volume Pumps • API 676 for Rotary Positive Displacement Pumps • ISO aimed at the medium duty single stage pumps ( Metric) • DIN. West German standard • VDMA West German standard for pump seals.

There are two problems with these standards:

• They were written for pumps equipped with jam packing. Most of the standards were written in the nineteen fifties at a time mechanical seals were not popular. In those days we had a lack of the modern materials that make mechanical seals practical. As an example Viton® was not invented until 1958 and did not come into general use until the sixties. Kalrez® did not come out until 1975 and in the eighties the duplex metals came into their own.

• The customer believes that by purchasing a standard design he is getting a good pump. Customers have the same problem with pump efficiency. They believe there is a correlation between efficiency and the quality of the pump, needless to say there is none! Problems caused by these standards are reflected in continual poor seal performance. The fact of the matter is that these standards reflect only an attempt to standardize envelope (outside) dimensions, nothing more!

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Unfortunately standardizing the length of pumps prevented manufacturers from designing short shafts that were not prone to the bending problems associated with low cost A.N.S.I. and I.S.O. design pumps, operating off of their best efficiency point (B.E.P).

Here is a list of some of the modifications you should make to your standard A.N.S.I. or I.S.O. pump if you want to get good mechanical seal and bearing life. Unless you are prepared to upgrade the pump seal and bearing life will always be less than desirable

WHAT TO MODIFY

• The stuffing box bore is too small for mechanical seals. In most cases there is not enough material to bore out so you will have to make or purchase a replacement part. Most of these standard stuffing boxes were designed for 3/8" or 10 mm. packing. You need at least 1" (25 mm.) radial clearance to take advantage of centrifugal force throwing solids away from the seal faces.

• When using mechanical seals install a recirculation line from the bottom of the stuffing box back to the suction of the pump. Try to tap the box as close to the face as possible to insure good circulation. Most quality seals come with this connection already installed in the gland.

• Because packing needs lubrication, the pump came equipped with a line from the discharge side to the stuffing box lantern ring connection. If you install a large sealing chamber in place of the narrow packing stuffing box that came as original equipment you should be able to eliminate almost all need for clean flushing liquid in the seal area. The only exception to this is if you are pumping a fluid close to its vaporization point. In that instance you do not want to lower stuffing box pressure because of the possibility of vaporizing the fluid in the stuffing box and possibly blowing open the seal faces

• Convert to Cartridge or Split Seals to insure correct seal installation and allow proper impeller settings in "Back Pull Out" or other types of pump designs.

• If you are using single stage centrifugal pumps convert to solid shafts with a low L3/D4 ratio to resist shaft bending. The back pull out design was made for easy sleeve removal. If you are using good mechanical seals, corrosion resistant shaft materials and labyrinth oil seals or positive bearing seals, there should be no need to replace pump shafts.

• Pump manufacturers are not required to provide L3/D4 ratio numbers that would predict shaft bending problems with their pump. The relationship between shaft size and shaft diameter is expressed in the ratio L3/D4. Try to keep it below 60 (2.5 Metric)

o "L" is the distance from the center of the inboard bearing to the center of the impeller (inches).

o "D" is the diameter of the shaft in the stuffing box area.(inches). • Substitute labyrinth or positive face seals for the lip or grease seals that are

installed in the bearing case. They will not only do a better job of keeping contaminants out of the bearing oil/grease, but they will not damage the

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expensive shaft. These seals also make sense in the motors to eliminate moisture from damaging the windings and contaminating the lubricating grease.

• Use only non- fretting mechanical seals. Shafts are too costly not to pay attention to this.

• The easiest way to get pump/motor alignment is with a "C" (inch) or "D" (metric) frame adapter. If you elect not to use the adapter you are in for a long process aligning the pump and driver correctly, and unless you are using split mechanical seals you are going to have to go through the procedure each time you change seals. You should be able to get the C or D frame adapter as part of your next power frame change or upgrade.

• Convert to a "Center Line" wet end if you are pumping liquids in excess of 200 degrees Fahrenheit (100 Centigrade) It will allow the suction flange to expand without causing pipe strain and wear ring damage.

• Do not use a vent on the bearing cavity of the pump. Each time the pump stops the vent will allow moisture to enter the bearing cavity as the oil cools down (this is called aspiration). You are much better off positively sealing the casing and installing an expansion chamber on the top of the casing to allow for air expansion.

• If you intend to use a closed impeller, end suction, centrifugal pump try to convert to a design that has adjustable wear rings

• Install a sight glass to be sure that the oil level is at the correct height. Too much oil is as bad as not enough. If you have a positive pressure oil mist system be sure that it does not vent to atmosphere. Oil mist systems require mechanical seals outboard of the bearings to prevent atmosphere contamination. If you have installed labyrinth seals, they will almost guarantee the correct oil level because excess oil will spill out of the labyrinth.

• Coat the inside of the bearing case with a suitable protective covering to prevent rusting and the leaching out of harmful substances from the bearing casting.

• Install magnetic plugs into the bottom of the bearing casing to attract loose metal shavings that would damage the bearings.

• Specify double volute designs any time the impeller diameter is 14" (356 mm.) or greater to prevent shaft deflection. Smaller size pumps do not lend themselves to this modification.

• Convert to a "Vortex" pump volute any time you are pumping liquid that contains lots of solids. Although the pump efficiency will be lower than conventional designs, the increased service life will more than compensate.

• At overhaul time substitute a medium or heavy weight power end for the light weight version that came with the pump and get most of the features we have discussed. Medium and heavy weight power ends are available for most popular pump brands.

In addition to these modifications mentioned here are some recommendations to help insure good seal and bearing life.

• Since a seal failure is the most common reason for shutting down a pump, Install a back up seal and convection tank to prevent unexpected shut downs.

• Change the bearing oil on a regular basis. Contact your favorite oil supplier for his recommendation and then follow his advice. If the inside of the bearing frame has been coated with a protective material to prevent rusting, avoid synthetic oils as their detergent action can damage these protective coatings.

• Maintain the proper oil level. Too much is just as bad as not enough. • Trim the impeller to obtain operation at the B.E.P. Throttling the pump discharge

is not the same thing. • If you are using open impellers, keep them adjusted to the correct clearance. • Install bearings by expanding the bore with an induction coil. Heating the bearing

in a pan of warm oil is not a good idea because the oil can easily be contaminated. • Install pressure gages on the suction and discharge of the pump. This is the only

way to tell if the pump is running near its B.E.P. • Do not specify Canned or Magnetic Pumps if the pumping fluid contains solids or

if it is a poor lubricant.

Pumps equipped with a "repeller" and some sort of static seal can usually be converted to a good mechanical seal. The problem with the repeller design is that in most of the designs the seal faces are designed to open when the pump is running and then close on any solids as the pump stops. The rule with mechanical seals is a simple one. "Keep the seal faces together" . Do not open them on purpose.

O.S.H.A. 1910 REGULATION

The regulation is predictably vague, and presently only applies to pressure vessels, storage tanks, processing piping, relief and vent systems, fire protection system components, emergency shut down systems, alarms, interlocks and the part that is important to you, pumps . For the first time Washington is telling the pump user that he has to now document the training he provides to those people (including contractors) that will be operating or repairing his pumps. Be sure to pick up a copy of this regulation for your library.

Here are some of the ingredients you will find in the regulation:

• The chemicals in the O.S.H.A. # 1910 specification are different than those chemicals identified by the E.P.A. for fugitive emission consideration. The O.S.H.A. list identifies those chemicals that are considered "extremely hazardous" chemicals. O.S.H.A. feels that the general industry standards are not sufficient for these chemicals

• Your employer is going to have to create a Process Safety Management audit team (PSM) that will audit company training programs along with insuring that present and future engineering practices conform to accepted standards and codes.

• The employer is going to have to identify the codes and standards he relied upon to establish his engineering practices. If he departs from these codes and standards, he must document that the design and construction are suitable for the intended purpose.

• The written training programs must be reviewed for adequacy of content, frequency of training, and the effectiveness of the training in terms of goals and objectives. These training programs must be revised if after the training the employee is not at the level of skill or knowledge that was expected.

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• Contract employees must also receive updated and current training. • If an accident occurs, the plant is going to have to prove that their training program was

adequate. • Any mechanical changes made by the maintenance department have to be evaluated to

determine whether operating procedures and practices also need to be changed. The term "Change" includes all modifications to equipment.

• For existing processes that have been shut down for turnaround or modification, the employer must ensure that any changes other than "replacement in kind" made to the process during shutdown go through the management of change procedures.

• Equipment installation jobs need to be properly inspected in the field for use of proper materials and procedures to insure that qualified workers do the job.

• The employer must ensure that the contractor has the appropriate job skills, knowledge, and certification.

• The regulations require detailed records of every action taken in maintaining or rebuilding a pump. The employer must identify which procedures were followed and why he elected to use those procedures. He must also identify the training that maintenance personnel had on repairing pumps in that service.

• Equipment used to process, store or handle hazardous chemicals has to be designed, constructed, installed and maintained to minimize the risk of release of such chemicals.

• The employer must prepare three lines of defense to prevent hazardous chemical from injuring personnel:

o Contain the chemical in the equipment. The use of two mechanical seals and a convection tank is a good example of containing the chemical.

o Control the release of the chemicals through venting with a seal quench and vent connection to a scrubber or flare, or to surge or overflow tanks designed to receive such chemicals. Dikes or designed drainage systems would be another alternative.

o A sensible evacuation system is the third line of defense.

If an accident happens and any of the listed chemicals are released to the environment, the employer is going to have to prove he did every thing he could have to prevent the accident and contain the spill. If O.S.H.A. does not agree with his assessment, the employer is likely to suffer stiff penalties.

Since you have knowledge that 90% of mechanical seals are failing prematurely (the carbon sacrificial face is not wearing out) I expect this new regulation should encourage your employer to send more people to seal and pump schools and enroll his engineering, maintenance, and supervisory people in an appropriate certification training program.

Why you must prime a centrifugal pump

Although the term "pressure" is not normally a part of a centrifugal pump man's vocabulary, we are going to have to discuss it for a couple of minutes.

The earth's atmosphere extends approximately fifty miles (80 Km.) above the earth, and rests on the earth with a weight equivalent to a layer of fresh water thirty four feet (10 meters) deep at sea level. To remove air from the pump cavities and the suction piping, the pump must develop enough head to equal the equivalent of this 14.7 psi., or one bar pressure. In an earlier paper we learned

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how to convert this height (head) to a pressure reading by use of the following formulas:

34 Ft./2.31 = 14.7 Psi.

10 Meters/10 = 1 Bar

Unlike a positive displacement pump that can pump a liquid to any head as long as the pump body is strong enough, and there is enough horsepower available, the centrifugal pump can only pump a liquid to its rated head. You will recall that this head was determined by, and limited to the diameter of the impeller and the impeller speed (rpm.)

Since the weight of water is approximately 8000 times that of air (50 miles vs. 34 feet or 80 Km. vs. 10 meters) the centrifugal pump can produce only 1/8000 of its rated liquid pressure. In other words, for every one foot water has to be raised to prime the pump, the centrifugal pump must produce a discharge head of approximately 8000 feet (each meter requires a head of 8000 meters) and that is impossible with conventional impeller diameters and speeds.

All of this means that if you intend to use a centrifugal pump you are going to have to come up with some sensible method of priming it. Your choices will include :

• Install a foot valve in the suction piping to insure the liquid will not drain from the pump casing and suction piping. Keep in mind that these valves have a nasty habit of leaking.

• Evacuate the air in the system with a positive displacement priming pump operating between the pump and a closed discharge valve.

• Fill the pump with liquid prior to starting it. • Convert the application to a self priming pump that maintains a reservoir of liquid

at its suction.

How efficient is your pump?

A few years ago, efficiency became "the name of the game". Automotive companies advertised "miles per gallon (liters per 100 kilometers) information in their advertisements and appliance manufacturers published kilowatt consumption numbers along with their pricing information.

Unfortunately high efficiency also means higher maintenance costs because you are required to maintain tighter tolerances and keep the flow passages smooth and free from obstructions. The demise of the double volute pump design in smaller size pumps, is a perfect example of the increase in mechanical seal problems as the efficiency of the volute pump was increased to satisfy consumer demand.

Maybe the "trade off" is acceptable as long as you are dealing with accurate numbers, but are you really doing that? Is the efficiency shown on the pump curve accurate? How was the data taken? What was included in the data, and more important, what was left out? As an example:

• Was the data generated on a dynamometer with a constant speed motor? • Are you going to run at the same speed as shown on the performance curve, or are

you running with an induction motor that slips 2% to 5% and you are not sure of the actual speed? Horse power (K.W.) varies as the cube of the change in speed at the best efficiency point, so a small variation in speed can make a big difference in efficiency.

• Was the published efficiency data generated with a seal or packing in the stuffing box? The type of packing or seal used can alter the load they consume.

• Was there an elbow at the suction of the pump? • Was the inside of the volute polished or coated with a low friction material when

the test was made? • How were the bearings lubricated, and were all of the losses considered in the

published numbers? • The final numbers will vary with the motor efficiency, and that will vary with the

load on the motor.

If you would like to keep the pump salesman honest, take the data from his pump curve and then make the following calculation:

In inch sizes : GPM x TDH / 3960 = WHP

• GPM = Gallon per minute at the best efficiency point • TDH = Total discharge head (measured in feet), as shown on the pump

curve&emdash; at the best efficiency point) • WHP = Water horse power, or the amount of horse power the pump is generating.

If we refer to the above pump curve, and insert the numbers into our formula, we would get:

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in inch size: GPM x TDH / 3960 = 250 X 300 / 3960 = 18.9

You then divide this number by the efficiency shown on the pump curve:

18.9 / .60 = 31.5 horsepower required to generate the WHP. If this number is lower than the horsepower shown on the performance curves, the efficiency date is questionable. As an example:

If the performance curve showed a requirement for 40 Horse power, the actual efficiency would be 18.9 water horse power40 pump horsepower = .47 or 47% actual efficiency.

Doing the same thing in the metric system we would get:

• M3/ HR = Cubic meters per hour of capacity as measured at the best efficiency point on the pump curve.

• TDH = Total discharge head, in meters, at the best efficiency point. • WKW = Water kilowatts of power being generated by the pump.

Referring to the above diagram, and putting in the numbers :

M3 / HR X TDH / 360 = 68 x 76 / 360 = 14.36 WKW. The curve shows a 60% efficiency so:

14.36 water kilowatts / 0.60 efficiency = 23.93 Kilowatts required. If this number is lower than shown on the pump performance curve, the efficiency of the plump is questionable. As an example:

If the pump performance curve showed a requirement for a 30 Kilowatt input, the actual efficiency would be:

14.36 water horse power / 30 Kilowatts required = 48 % actual efficiency.

The fact of the matter is that you seldom operate at the best efficiency point so the numbers become even more depressing. The point is that efficiency should only be one of the points taken into consideration when you purchase a centrifugal pump of a given head, material and capacity. Equally if not more important should be:

• The L3/D4 number of the shaft. Is the number below 60 in inch sizes or 2 in metric? • What kind of mechanical seal is installed? Will it seal fugitive emissions? • How are the bearings being lubricated? • How are the bearings sealed? Will the seal damage the expensive shaft?

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• How is the thrust bearing being retained? In operation the impeller thrusts towards the volute. Are you relying upon a simple snap ring?

• Is the pump a centerline design? It should be if the product you are pumping is greater than 200°F (100 C.)

• Is the bearing case vented to atmosphere? If it is, it will allow moisture to penetrate when the pump stops.

• Has a "C" or "D" frame adapter been installed to reduce alignment time? • Can the wear rings or open impeller be easily adjusted to compensate for normal

wear so that you can keep the efficiency you paid for? • Can the seal compensate for thermal growth, or impeller adjustment?

You can save money by lowering operating costs (efficiency) or increasing the time between repairs (design). Be sure you consider both when you make your pump buying decision.


When we talk about automobiles and discuss efficiency, we mean how many miles per gallon, or liters per 100 kilometers. When we discuss centrifugal pumps we are comparing the amount of work or power we get out of the pump to the amount of power we are putting into the pump. As an example:

How do we measure the horsepower or kilowatts coming out of the pump? All we have to do is multiply the pump head by the weight of the liquid being pumped, and then use a simple conversion number. Let's take an example:

Flow = 300 gallons per minute of fresh water as measured coming from the pump discharge.

Head = 160 feet. We measured it at the discharge side of the pump and corrected it for the fact that the gage was two feet above the pump center line. Look at the following diagram where we have calculated the discharge head from the formula shown on the right hand side of the illustration. If there were any positive head on the suction side of the pump that head would have to be subtracted. A negative suction head would be added to the discharge head.

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The centrifugal pump pumps the difference between the suction and the discharge heads. There are three kinds of discharge head:

• Static head. The height we are pumping to, or the height to the discharge piping outlet that is filling the tank from the top. Note: that if you are filling the tank from the bottom, the static head will be constantly changing.

• Pressure head. If we are pumping to a pressurized vessel (like a boiler) we must convert the pressure units (psi. or Kg.) to head units (feet or meters).

• System or dynamic head. Caused by friction in the pipes, fittings, and system components. We get this number by making the calculations from published charts ( non included in this paper, but available in the chart section of this web site).

Suction head is measured the same way.

• If the liquid level is above the pump center line, that level is a positive suction head. If the pump is lifting a liquid level from below its center line, it is a negative suction head.

• If the pump is pumping liquid from a pressurized vessel, you must convert this pressure to a positive suction head. A vacuum in the tank would be converted to a negative suction head.

• Friction in the pipes, fittings, and associated hardware is a negative suction head. • Negative suction heads are added to the pump discharge head, positive suctions

heads are subtracted from the pump discharge head.

Here is the formula for measuring the horsepower out of the pump:

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Remember that we are using the actual horsepower or kilowatts going into the pump and not the horsepower or kilowatts required by the electric motor. Most motors run some where near 85% efficient.

An 85% efficient motor turning a 76% efficient pump, gives you a real efficiency of 0 .85 x 0.76 = 0 .65 or 65% efficient.

A survey of popular pump brands demonstrates that pump efficiencies range from 15% to over 90%. The question then arises, "Is this very wide range due to poor selection, poor design, or some other variable which would interfere with good performance?" The best available evidence suggests that pump efficiency is directly related to " the specific speed number " with efficiencies dropping dramatically below a number of 1000 . Testing also shows that smaller capacity pumps exhibit lower efficiencies than higher capacity designs.

Now that we have learned that pump efficiency is closely related to the shape of the impeller, and the impeller shape is usually dictated by the operating conditions, you should be aware of various conditions that decrease the efficiency of your pump. These would include:

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• Packing generates approximately six times as much heat as a balanced mechanical seal.

• Wear rings and impeller clearances are critical. Anything that causes these tolerances to open will cause internal recirculation that is wasting power as the fluid is returned to the suction of the pump. If the wear ring is rubbing, the generated heat is consuming power.

• A bypass line installed from the discharge side of the pump to the suction piping. The heat generated from this recirculation can, in some cases, cause pump cavitation as it heats the incoming liquid.

• A double volute design pump restricts the discharge passage lowering the overall efficiency.

• Running the pump with a throttled discharge valve. • Eroded or corroded internal pump passages will cause fluid turbulence. • Any restrictions in the pump or piping passages such as product build up, a

foreign object, or a stuck check valve. • Over lubricated or over loaded bearings. • Rubbing is a major cause. It can be caused by:

o Misalignment between the pump and driver. o Pipe strain. o Impeller imbalance. o A bent shaft. o A close fitting bushing. o Loose hardware. o A protruding gasket rubbing against the mechanical seal. o Cavitation. (5 kinds) o Harmonic vibration. o Improper assembly of the bearings, seal, wear rings, packing, lip seals etc.. o Thermal expansion of various components in high temperature

applications. The impeller can hit the volute, the wear rings can come into physical contact etc.

o Solids rubbing against the rotating components, especially the seal. o Operating too far off of the best efficiency point of the pump. o Water hammer and pressure surges. o Operating at a critical speed. o Dynamic, non O-ring elastomers that cannot flex and roll, but must slide,

eventually fretting the shaft or sleeve. o A build up of product on the inside of the stuffing box rubbing against the

mechanical seal. o Grease or lip seals rubbing the shaft next to the bearings. o Over tightening packing or improper seal installation.

• Vortex pumps can lower efficiency by as much as 50%.

All about specific speed

Specific speed is a term used to describe the geometry (shape) of a pump impeller. People responsible for the selection of the proper pump, for their application, can use this Specific Speed information to :

• Select the shape of the pump curve. • Determine the efficiency of the pump. • Anticipate motor overloading problems. • Predict N.P.S.H. requirements. • Select the lowest cost pump for their application.

Specific speed is defined as "the speed of an ideal pump geometrically similar to the actual pump, which when running at this speed will raise a unit of volume, in a unit of time through a unit of head".

The performance of a centrifugal pump is expressed in terms of pump speed, total head, and required flow. This information is available from the pump manufacturer's published curves. Specific speed is calculated from the following formula, using data from these curves at the pump's best efficiency point (B.E.P.):

N = The speed of the pump in revolutions per minute (rpm.)

Q = The flow rate in liters per minute ( for either single or double suction impellers)

H = The total dynamic head in meters

Please refer to the following chart:

Pumps are traditionally divided into three types: radial flow, mixed flow, and axial flow. When you look at the above chart you can see there is a gradual change

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from the radial flow impeller, which develops pressure principally by the action of centrifugal force, to the axial flow impeller, which develops most of its head by the propelling or lifting action of the vanes on the liquid.

In the specific speed range of approximately 1000 to 6000 double suction impeller are used as frequently as the single suction impellers.

If you substitute other units for flow and head the numerical value of Ns will vary. The speed is always given in revolutions per minute (rpm.). Here is how to alter the Specific Speed number (Ns) if you use other units for capacity and head :

• United States ....Q = G.P.M. and H = feet. Divide the Ns by 1.63 • British ............Q = Imp.G.P.M. and H = feet. Divide the Ns by 1.9 • Metric ............Q = M3/hour and H = meters. Divide the Ns by 1.5

As an example we will make a calculation of Ns in both metric and U.S. units :

• Q= 110 L/sec. or 396 M3/ hour or 1744 G.P.M. • H = 95 meters or 312 feet • Speed = 1450 rpm.

If the above results were describing an actual application, we would notice that it was a low specific speed, radial flow pump, meaning It would be a large pump with a low efficiency. Going to 2900 rpm. or higher would increase the Ns to 1000 or more, meaning a smaller pump with a much higher efficiency, but this higher rpm. would have other possible consequences :

• The higher efficiency would allow you to use a less powerful driver that would reduce your operating costs.

• A smaller pump makes associated hardware cheaper. For instance, a smaller diameter shaft means a lower cost mechanical seal and lower cost bearings.

• Cavitation could become a problem as the increase in speed means an increase in the N.P.S.H. required.

• If you are pumping an abrasive fluid, abrasive wear and erosion will increase with increasing speed.

• Many single mechanical seals have problems passing fugitive emission standards at the higher pump speeds.

• High heat is a major cause of bearing failure. The higher pump speeds contribute to the problem.

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The following diagram illustrates the relationship between specific speed and pump efficiency. In general, the efficiency increases as Ns increases.

Specific speed also relates to the shape of the individual pump curve as it describes head, capacity, power consumption and efficiency.

In the above diagram you will note that :

• The steepness of the head/ capacity curve increases as specific speed increases. • At low specific speed, power consumption is lowest at shut off and rises as flow

increases. This means that the motor could be over loaded at the higher flow rates unless this was considered at the time of purchase.

• At medium specific speed the power curve peaks at approximately the best efficiency point. This is a non overloading feature meaning that the pump can work safely over most of the fluid range with a motor speed to meet the B.E.P. requirement.

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• High specific speed pumps have a falling power curve with maximum power occurring at minimum flow. These pumps should never be started with the discharge valve shut. If throttling is required a motor of greater power will be necessary.

Keep in mind that efficiency and power consumption were calculated at the best efficiency point (B.E.P.). In practice most pumps operate in a throttled condition because the pump was oversized at the time it was purchased. Lower specific speed pumps may have lower efficiency at the B.E.P., but at the same time will have lower power consumption at reduced flow than many of the higher specific speed designs.

The result is that it might prove to be more economical to select a lower specific speed design if the pump had to operate over a broad range of capacity.

The oversized pump

Do a survey of any process plant and you will find that a high percentage of the centrifugal pumps are oversized. There must be a reason why this is such a common problem, so here are a few of them :

• Safety margins were added to the original calculations. Several people are involved in the pump buying decision and each of them is afraid of recommending a pump that proves to be to small for the job.

• It was anticipated that a larger pump would be needed in the future, so it was purchased now to save buying the larger pump later on.

• It was the only pump the dealer had in stock and you needed one badly. He might have offered you a "special deal" to take the larger size.

• You took the pump out of your spare parts inventory. Capital equipment money is scarce so the larger pump appeared to be your only choice.

• You purchased the same size pump as the one that came out of the application and that one was over sized also.

Obviously this larger pump and motor required a higher investment, but since we are not using the full power are we really paying too much for the daily operation? The easiest way to find the answer to this question is to look at a typical pump curve and make our calculations from the numbers we get.

You can use any of the following formulas to make your calculations:

Here is as typical pump curve. It can be used for both inch and metric examples.

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Let us assume that the application requires a pump that moves the liquid at :

300 gpm. to a 156 foot head with an efficiency rating of 60%

156 x 300 / 5308 = 8.8 Kilowatts being produced, and 8.8 / 0.60 efficiency = 14.7 Kilowatts required

As shown in the above drawing, we should be using impeller "E" to do this, but we have an oversized pump so we are using the larger impeller "A" with the pump discharge valve throttled back to 300 gpm. giving us an actual head of 250 feet and a 50% efficiency. Now our Kilowatts look like this:

250 x 300 / 5308 = 14.1 KW being produced, and 14.1 / 0.50 efficiency = 28.2 KW

required to do this. If 28.2 KW is being used and only 14.7 KW are required, it means that we are paying for an extra 13.5 KW to pump against the throttled discharge valve.

If this pump runs 24 hours per day that would be 8760 hours this year, and at a power cost of $0.05 cents per Kilowatt hour it would cost your company an additional:

8760 hours. x .05 cents per Kilowatt hour x 13.5 Kilowatts = $5913.00 per year, extra operating cost.

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Now we will work the same problem in the metric system:

Assume that we need to pump 68 m3/hr. to a 47 meter head with a pump that is 60% efficient at that point.

68 x 47360 = 8.9 Kilowatts being produce, and 8.9 / 0.60 efficient = 14.8 Kilowatts required to do this.

As shown in the drawing, we should be using impeller "E" to do this, but we have an oversized pump so we are using the larger impeller "A" with the pump discharge valve throttled back to 68 cubic meters per hour, giving us an actual head of 76 meters. Now our Kilowatts look like this:

68 x 76360 = 14.3 Kilowatts being produced by the pump, and 14.3 / 0.50 efficient = 28.6 Kilowatts required to do this.

Subtracting the amount of kilowatts we should have been using gives us:

28.6 - 14.8 = 13.8 extra kilowatts being used to pump against the throttled discharge valve. If the pump runs twenty four hours a day that would be 8760 hours per year, times 13.8 extra kilowatts equals 120,880 kw. Multiply this number by how much you spend for a kilowatt hour of electricity and you will see that the over sized pump is costing you a lot of money. In this example the extra cost of the electricity could almost equal the cost of purchasing the pump.

How to read a pump curve

Please look at the above illustration. You will note that I have plotted the head of the pump against its capacity. The head of a pump is read in feet or meters. The capacity units will be either gallons per minute, liters per minute, or cubic meters per hour.

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According to the above illustration this pump will pump a 40 capacity to about a 110 head, or a 70 capacity to approximately a 85 head (you can substitute either metric or imperial units as you see fit)

The maximum head of this pump is 115 units. This is called the maximum shutoff head of the pump. Also note that the best efficiency point (BEP) of this impeller is between 80% and 85% of the shutoff head. This 80% to 85% is typical of centrifugal pumps, but if you want to know the exact best efficiency point you must refer to the manufacturers pump curve.

Ideally a pump would run at its best efficiency point all of the time, but we seldom hit ideal conditions. As you move away from the BEP the shaft will deflect and the pump will experience some vibration. You will have to check with your pump manufacturer to see how far you can safely deviate from the BEP (a maximum of 10% either side is typical)

Now look at the following illustration:

Note that I have added some additional curves to the original illustration. These curves show what happens when you change the diameter of the impeller.

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Impeller diameter is measured in either inches or millimeters. If we wanted to pump at the best efficiency point with a 11.5 impeller we would have to pump a capacity of 50 to a 75 head.

The bottom half of the illustration shows the power consumption at various capacities and impeller diameters. I have labeled the power consumption horsepower, but in the metric system it would be called kilowatts

Each of the lines represents an impeller diameter. The top line would be for the 13 impeller the second for the 12.5 etc. If we were pumping a capacity of 70 with a 13 impeller it would take about 35 horsepower. A capacity of 60 with the 12 impeller would take about 20 horsepower.

Most pump curves would show you the percent of efficiency at the best efficiency point . The number varies with impeller design and numbers from 60% to 80% are normal.

When you will look at an actual pump curve you should have no trouble reading the various heads and corresponding capacities for the different size impellers. You will note however, that the curve will usually show an additional piece of information and that is NPSHR which stands for net positive suction head required to prevent the pump from cavitating.

Depending upon the pump curve you might find a 10 foot (3.0 meter) NPSH required head at a capacity of 480 Gallons per minute (110 cubic meters per hour) if you were using a 13 inch (330 mm.) diameter impeller.

You should keep in mind that the manufacture assumed you were pumping 20° C ( 68° F ) fresh water and the N.P.S.H. Required was tested using this assumption. If you are pumping water at a different temperature or if you are pumping a different fluid, you are going to have to add the vapor pressure of that product to the N.P.S.H. Required. The rule is that Net Positive Suction Head Available minus the Vapor Pressure of the product you are pumping (converted to head) must be equal to or greater than Net Positive Suction Head Required by the manufacturer.

Suppose we wanted to pump some liquid Butane at 32 degrees Fahrenheit (0 degrees Centigrade) with this pump. If we look at the curve for Butane on a vapor pressure chart similar to the one shown in the charts and graphs section of this web site you will note that Butane at 32°F needs at least 15 psi (1,0 Bar) to stay in a liquid state. To convert this pressure to head we use the standard formula :

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In other words Butane at this temperature would not vaporize as long as I had the above absolute heads available at the suction side of the pump.

Understanding the system curve

Every pump manufacturer would like to recommend the perfect pump for your application. To do this he would like you to provide him with an accurate system curve that would describe the capacity and head needed for your various operating conditions. Once he has your system curve, he can plot his pump curves on top of the system curve and hopefully select something that will come close to your needs. Without this system curve, neither one of you has much of a chance of coming up with the right pump.

To create a system curve we plot the desired capacities against the required head over the total anticipated operating range of the pump. The head will be measured in feet or meters and the capacity will be measured in gallons per minute or cubic meters per hour.

Some of the confusion begins when we realize that there are three different kinds of head:

STATIC HEAD This is the vertical distance measured from the center line of the pump to the height of the piping discharge inside the tank. Look at figure "A" and note that the piping discharge is below the maximum elevation of the piping system. We do not use the maximum elevation in our calculations because the siphoning action will carry the fluid over this point once the piping is full of liquid. This is the same action that lets you siphon gasoline out of an automobile to a storage can.

The pump will have to develop enough head to fill the pipe and then the siphoning action will take over. The pump operating point should move back towards the best efficiency point (B.E.P.) if the pump was selected correctly.

FIGURE "A"

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DYNAMIC OR SYSTEM HEAD As the liquid flows through the piping and fittings, it is subject to the friction caused by the piping inside finish, restricted passages in the fittings and hardware that has been installed in the system. The resulting "pressure drop" is described as a "loss of head" in the system, and can be calculated from graphs and charts provided by the pump and piping manufacturers. These charts are not included with this paper, you can find them in the Hydraulic Institute Manuals. This "head" loss is related to the condition of the system and makes the calculations difficult when you realize that older systems may have "product build up" on the piping walls, filters, strainers, valves, elbows, heat exchangers, etc., making the published numbers some what inaccurate.

A general "rule of thumb" says that the friction loss in clean piping will vary approximately with 90% of the square of the change in flow in the piping, and 100% of the square with the change of flow in the fittings and accessories. You calculate the change in flow by dividing the new flow by the old flow and then square the number. As an example:

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In the original application system, loss was a combination of the loss through the piping and the loss through the fittings for a total of 100 feet at 200 gallons per minute. When we increased the flow to 300 gallons per minute our system head changed to a total of 208.13 feet. This change would have to be added to the static and pressure heads to calculate the total head required for the new pump.

Please note that the pump is pumping the difference between the suction head and the discharge head, so if you fail to consider that the suction head will be either added to or subtracted from the discharge head, you will make an error in your calculations. The suction head will be negative if you are lifting liquid from below ground or if you are pumping from a vacuum. It will be positive if you are pumping from a tank located above ground. If the suction head is pressurized, this pressure must be converted to head and subtracted from the total head required by the pump.

A centrifugal pump will create a head/capacity curve that will generally resemble one of the curves described in figure "B" The shape of the curve is determined by the Specific Speed number of the impeller.

Centrifugal pumps always pump somewhere on their curve, but should be selected to pump as close to the best efficiency point (B.E.P.) as possible. The B.E.P. will fall some where between 80% and 85% of the shut off head (maximum head).

The manufacturer generated these curves at a specific R.P.M.. Unless you are using synchronous motors (you probably are using induction motors on your pumps) you will have to adjust the curves to match your actual pump speed. Put a tachometer on the running motor and record the rpm. difference between your pump and the speed shown on the pump manufacturer's published curve. You can use the pump affinity laws to approximate the change.

POSITIVE DISPLACEMENT PUMPS have a different shaped curve. They look something like Figure "C".

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In this system, the head remains a constant as the capacity varies. This is a typical application for:

• A boiler feed pump that is supplying a constant pressure boiler with a varying steam demand. This is a very common application in many process systems or aboard a ship that is frequently changing speeds (answering bells).

• Filling a tank from the top and varying the amount of liquid being pumped, is the normal routine in most process plants. The curve will look like this if the majority of the head is either static or pressure head.

The second system is the ideal one, Figure "E" describes it:

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In this system the entire head is system head so it will vary with the capacity. Look for this type of curve in the following applications:

• A circulating hot or cold water heating/ cooling system. • Pumping to a non pressurized tank, a long distance from the source with little to

no elevation involved. Filling tank cars is a typical application.

System curve "G" is a common one. It is a combination of static, pressure and system heads.

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Once the pump manufacturer has a clear idea as to the shape of your system curve, and the head and capacity numbers needed he can then select the proper centrifugal pump. The shape of his curve will be pretty much determined by the specific speed number of the impeller.

In addition to specific speed he can select impeller diameter, impeller width, pump rpm., and he also has the option of series or parallel operation along with the possibility of using a multi-stage pump to satisfy your needs.

The sad fact is that most pumps are selected poorly because of the desire to offer the customer the lowest possible price. A robust pump, with a low L3/D4, is still your best protection against seal and bearing premature failure when the pump is operating off of its best efficiency point. Keep the following in mind as you select your pump:

• A centrifugal pump will pump where the pump curve intersects the system curve. This may bear no relationship to the best efficiency point (B.E.P.), or your desire for the pump to perform a specific task.

• The further off the B.E.P. you go, the more robust the pump you will need. This is especially true if you have replaced the packing with a mechanical seal and no longer have the packing to act as a support bearing when the shaft deflects. Shaft deflection is always a major problem at start up.

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• When you connect pumps in parallel, you add the capacities together. The capacity of a pump is determined by the impeller width and r.p.m.. The head of a centrifugal pump is determined by the impeller diameter and rpm. If the heads are different, the stronger pump will throttle the weaker one, so the impeller diameters and rpm's must be the same if you connect pumps in parallel. Check the rpm's on these pumps if you are experiencing any difficulties.

• If you connect the pumps in series, the heads will add together, so the capacities must be the same or one of them will probably cavitate. You could also have a problem operating too far to the right of the best efficiency point with a possible motor "burn out".

• When you vary the speed of a centrifugal pump, the best efficiency point comes down at an angle. The affect is almost the same as changing the diameter of the impeller. This means that the variable speed motor will work best on a system curve that is exponential (Figure "F"). Unfortunately most process and boiler feed pump system curves are not exponential.

• Pump curves are based on a speed of 1750, 3500, 1450, or 2900 r.p.m.. Electric induction motors seldom run at these speeds because of "slip". You can estimate that a 2% to a 5% slip is normal in these pumps with the "slip" directly related to the price of the motor.

• You should also keep in mind that if the motor is running at its best efficiency point that does not mean that the pump is running at its B.E.P..

Since you will be using pumps that were supplied at the lowest cost, you can do the following to resist some of the shaft displacement:

• Use a solid shaft. Sleeves often raise the L3/D4 number to over 60 (2 in the metric system), and this is too high a number for reliable seal performance.

• Try to keep the mechanical seal as close to the bearings as possible. It is the mechanical seal that is the most sensitive to shaft deflection and vibration.

• Once the seal has been moved closer to the bearings, you can install a sleeve bearing in the packing space to support the shaft when the pump is operated off of its B.E.P. This is especially important at start up, or any time a pump discharge valve is operated.

• Stop the cavitation if you are experiencing any. • Balance the rotating assembly. • Check that the shaft is not bent or the rotating assembly is not out of dynamic

balance. • Use a "C" or "D" frame adapter to solve pump- motor alignment difficulties. • A center line design wet end can be used if pipe strain, due to temperature

expansion, is causing an alignment problem.

Do not trust the system prints to make your calculations. The actual system always differs from that shown on the print, because people tap into the lines, using the pumped fluid for a variety of purposes and after having done so forget to change or "mark up" the original system print. You are going to have to "walk down" the system and note the pipe length, the number of fittings, etc., to make

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an accurate system head calculation. Do not be surprised to find that the discharge of your pump is hooked up to the discharge of another pump further down the line. In other words, the pumps are connected in parallel and no body knows it. Pressure recorders (not gauges) installed at the pump suction and discharge is another technique you can use to get a better picture of the system or dynamic head. They will show you how the head is varying with changes in flow.

Pump selection is simple but not easy. Do not depend upon the knowledge of the local pump salesman to select the correct pump for you. In many cases he is prepared to sell his pump at cost&emdash;to get the spare parts business. If you are purchasing pumps at too big a discount&emdash;something is wrong, there is no free lunch. Keep in mind that if several people are involved in the selection process each of them will commonly add a safety factor to the calculated pump size. These factors added together can cause you to purchase a pump that is very much over sized.

Calculating the total system head in USCS units

USCS stands for "United States Customary System Units" as opposed to the SI (Le Syst`eme International d`Units) or metric units that have been adopted by the International standards Organization (ISO). In a future paper I will present another paper using the metric units, but for the moment it is not convenient to present it in both systems.

It turn out that "head" is a very convenient term in the pumping business. Capacity is measured in gallons per minute, and each gallon of liquid has weight, so we can easily calculate the pounds per minute being pumped. Head or height is measure in feet, so if we multiply these two together we get foot- pounds per minute which converts directly to work at the rate of 33,000 foot pounds per minute equals one horsepower.

Pressure is not as convenient a term because the amount of pressure that the pump will deliver depends upon the weight (specific gravity) of the liquid being pumped and the specific gravity changes with temperature, fluid, and fluid concentration.

If you will refer to FIG 1, you should get a clear picture of what is meant by static head. Note that we always measure from the center line of the pump to the highest liquid level

To calculate head accurately we must calculate the total head on both the suction and discharge sides of the pump. In addition to the static head we will learn that there is a head caused by resistance in the piping, fittings and valves called friction head, and a head caused by any pressure that might be acting on the liquid in the tanks including atmospheric pressure, called " surface pressure head".

Once we know these heads it gets simple, we will then subtract the suction head from the discharge head and the amount remaining will be the amount of head that the pump must be able to generate at the rated flow. Here is how it looks in a formula:

System head = total discharge head - total suction head

H = hd - hs

The total discharge head is made from three separate heads:

hd = hsd + hpd + hfd

• hd = total discharge head • hsd = discharge static head • hpd = discharge surface pressure head • hfd = discharge friction head

The total suction head also consists of three separate heads

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hs = hss + hps - hfs

• hs = total suction head • hss = suction static head • hps = suction surface pressure head • hfs = suction friction head

As we make these calculations, you must sure that all calculations are made in either "feet of liquid gauge" or "feet of liquid absolute". In case you have forgotten "absolute means that you have added atmospheric pressure (head) to the gauge reading.

Now we will make some actual calculations:

Figure #2 demonstrates that the discharge head is still measured to the liquid level, but you will note that it is below the maximum height of the piping.

Although the pump must deliver enough head to get up to this maximum piping height, it will not have to continue to deliver this head when the pump is running because of the "siphon effect". There is of course a maximum siphon effect. It is derived from: 14.7 psi (atmospheric pressure) x 2.31 feet / psi = 33.4 feet maximum siphon effect.

We will begin with the total suction head calculation

1. The suction head is negative because the liquid level in the suction tank is below the centerline of the pump:

hss = - 6 feet

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2. The suction tank is open, so the suction surface pressure equals atmospheric pressure :

hps = 0 feet gauge

3. You will not have to calculate the suction friction head, I will tell you it is:

hfs = 4 feet at rated flow

4. The total suction head is a gauge value because atmosphere was given as 0,

hs = hss + hps - hfs = -6 +0 -4 = -10 feet of liquid gauge at rated flow

The total discharge head calculation

1. The static discharge head is:

hsd = 125 feet

2. The discharge tank is also open to atmospheric pressure, thus:

hpd = 0 feet, gauge

3. I will give you the discharge friction head as:

hfd = 25 feet at rated flow

4. The total discharge head is:

hd = hsd + hpd + hfd = 125 + 0 + 25 = 150 feet of liquid gauge at rated flow

The total system head calculation:

H = hd - hs = 150 - (-10)= 160 feet of liquid at rated flow

Note: did you notice that when we subtracted a minus number (-10) from a positive number (150) we ended up with a positive 160 because whenever you subtract minus numbers it is the same as adding them? If you have trouble with this concept you can learn more about it from a mathematics book.

Our next example involves a few more calculations, but you should be able to handle them. In this example we are going to learn how to handle a vacuum application. Pipe friction numbers are taken from the Hydraulic Institute Engineering Data Book. You can get a copy of this publication from your library if you want to see the actual charts. I have some of this information in the chart section of this web site.

Specifications:

1. Transferring 1000 gpm. weak acid from the vacuum receiver to the storage tank

2. Specific Gravity - 0.98

3. Viscosity -equal to water

4. Piping - All 6" Schedule 40 steel pipe

5. Discharge piping rises 40 feet vertically above the pump centerline and then runs 400 feet horizontally. There is one 90° flanged elbow in this line

6. Suction piping has a square edge inlet, four feet of pipe, one gate valve, and one 90° flanged elbow all of which are 6" in diameter.

7. The minimum level in the vacuum receiver is 5 feet above the pump centerline.

8. The pressure on top of the liquid in the vacuum receiver is 20 inches of mercury, vacuum.

To calculate suction surface pressure use one of the following formulas:

• inches of mercury X 1.133specific gravity = feet of liquid • pounds per square inch X 2.31specific gravity = feet of liquid

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• Millimeters of mercury X 122.4 x specific gravity = feet of liquid

Now that you have all of the necessary information we will begin by dividing the system into two different sections, using the pump as the dividing line.

Total suction head calculation

1. The suction side of the system shows a minimum static head of 5 feet above suction centerline. Therefore, the static suction head is:

hss = 5 feet

2. Using the first conversion formula, the suction surface pressure is:

hps = -20 Hg X 1.133/ 0.98 = -23.12 feet gauge

3. The suction friction head, fs, equals the sum of all the friction losses in the suction line. Friction loss in 6" pipe at 1000 gpm from table 15 of the Hydraulic Institute Engineering Data Book, is 6.17 feet per 100 feet of pipe.

in 4 feet of pipe friction loss = 4/100 x 6.17 = 0.3 feet

Friction loss coefficients (K factors) for the inlet, elbow and valve can be added together and multiplied by the velocity head:

FITTING K FROM TABLE6" Square edge inlet 0.50 32 (a) 6" 90 flanged elbow 0.29 32 (a) 6" Gate valve 0.11 32 (b)

Total coefficient, K = 0.90

Total friction loss on the suction side is:

hfs = 0.3 + 1.7 = 2.0 feet at 1000 gpm.

4. The total suction head then becomes:

hs = hss + hps - hfs = 5 + (-23.12) - 2.0 = -20.12 feet, gauge at 1000 gpm.

Total discharge head calculation

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1. Static discharge head = hsd = 40 feet

2. Discharge surface pressure = hpd = 0 feet gauge

3. Discharge friction head = hfd = sum of the following losses :

Friction loss in 6" pipe at 1000 gpm. from table 15, is 6.17 feet per hundred feet of pipe.

In 440 feet of pipe the friction loss = 440/100 x 6.17 = 27.2 feet

Friction loss in 6" elbow:

from table 32 (a), K = 0,29

from table 15, V2/2g = 1.92 at 1000 gpm.

Friction loss = K V2/2g = 0.29 x 1.92 = 0.6 feet

The friction loss in the sudden enlargement at the end of the discharge line is called the exit loss. In systems of this type where the area of the discharge tank is very large in comparison to the area of the discharge pipe, the loss equals V2/2g, as shown in table 32 (b).

Friction loss at exit = V2/2g = 1.9 feet

The discharge friction head is the sum of the above losses, that is:

hfd = 27.2 + 0.6 + 1.9 = 29.7 feet at 1000 gpm.

4. The total discharge head then becomes:

hd = hsd + hpd + hfd = 40 + 0 + 29.7 = 69.7 feet, gauge at 1000 gpm.

c. Total system head calculation:

H = hd - hs = 69.7 - (-20.2) = 89.9 feet at 1000 gpm.

Our next example will be the same as the one we just finished except. that there is an additional 10 feet of pipe and another 90° flanged elbow in the vertical leg. The total suction head will be the same as in the previous example. Take a look at figure # 4

Nothing has changed on the suction side of the pump so the total suction head will remain the same:

hs = -20.12 feet, gauge at 100 gpm.


1. The static discharge head "hsd" will change from 40 feet to 30 feet, since the highest liquid surface in the discharge is now only 30 feet above the pump centerline.(This value is based on the assumption that the vertical leg in the discharge tank is full of liquid and that as this liquid falls it will tend to pull the liquid up and over the loop in the pipe line. This arrangement is called a siphon leg).

2. The discharge surface pressure is unchanged:

hpd = 0 feet

3. The friction loss in the discharge pipe will be increased by the additional 10 feet of pipe and the additional elbow.


The friction loss in the additional elbow = 0.6 feet

The friction head will then increase as follows:

hfd = 29.7 + 0.6 + 0.6 = 30.9 feet at 1000 gpm.

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The total discharge head becomes:


= 30 + 0 + 30.9

= 60.9 feet, gauge at 1000 gpm.

5. Total system head calculation

H = hd - hs = 60.9 - (-20.12) = 81 feet at 1000 gpm.

For our last example we will look at gauges. Take a look at FIG 5:

Specifications:

• Capacity - 300 gpm. • Specific gravity - 1.3 • Viscosity - Similar to water • Piping - 3 inch suction, 2 inch discharge • Atmospheric pressure - 14.7 psi.

Divide the heads into two sections again:

The discharge gauge head corrected to the centerline of the pump, in feet of liquid absolute is found by adding the atmospheric pressure to the gauge reading to get absolute pressure, and then converting to absolute head:

hdg = (130 + 14.7) x 2.31 / (1.3 Specific Gravity + 4 ) = 261.1 feet, absolute

Note the 4 foot head correction to the pump centerline.

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The discharge velocity head at 300 gpm. is found in table 9 of the Hydraulic Institute Engineering Data Book

hvd = 12.8 feet at 300 gpm.

The suction gauge reading is in absolute terms so it needs only to be converted to feet of liquid, absolute.

hgs = 40 x 2.3 / 11.3 +2 = 73.08 feet absolute


The suction velocity head at 300 gpm. is found in table 11 of the Pipe Friction Manual:

hvs = 2.6 feet at 300 gpm.

The total head developed by the pump is:

H = (hgd + hvd ) - ( hgs + hvs ) = (261.1 + 12.8) - (73.08 + 2.6)= 198.22 feet absolute at 300 gpm.

Estimating the shutoff head of a centrifugal pump:

In the fifteenth century the Swiss scientist Daniel Bernoulli learned that the combination of head and velocity was a constant throughout a piping system. He then wrote the formula showing the relationship between this liquid velocity, and resultant head. As many of you know, I often quote this formula in my pump and seal schools. The formula looks like this:

• V = Velocity or speed of the liquid at the impeller outside diameter (ft/sec. or meters/sec.)

• g = gravity = 32.2 feet / second2 or 9.8 meters / second2

My students have heard me quote this formula as the basis for my statement that you can estimate the shut off head of a 1750 rpm. centrifugal pump by squaring the diameter of the impeller. How did I come to that conclusion ? Lets look at the formula again, and we will start by defining velocity:

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Velocity is a measurement of speed using distance and time as the variables. The terms we use to discuss velocity are feet/second or meters/ second. In the inch system the velocity of the impeller outside diameter is determined by the following formula:

• d = diameter of the impeller • • rpm = speed of the impeller outside diameter • 12 = twelve inches in a foot • 60 sixty seconds in a minute

Now we will solve the formula. Substituting 1750 for the rpm we would get:

Going back to the original formula we will substitute the new value for "V"

This means that at 1750 rpm the shutoff head is 90% of the diameter of the impeller squared

If you will check a typical pump curve as supplied by the pump manufacturers, you will learn that the shut off head actually varies from 90% to 110% of the diameter of the impeller squared. I elected to use 100% because it is a sensible average and in some cases it accounts for the additional velocity added to the fluid as it moves from the impeller eye to the impeller outside diameter.

If we substitute 3500 rpm for the speed, the new numbers would look like this


We can round out the 3.6 to 4.0 and say that at 3500 rpm the shutoff head equals approximately the outside diameter of the impeller squared, times four.

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39

se decimeters instead. It will make the calculations a lot simpler because you will be

It is a little trickier in the metric system. Instead of using millimeters when measuring the impeller diameter, move over two decimal places and u

using more convenient, larger numbers.

Inserting the numbers into the formula we would get a velocity of:

Going back to the head formula we would get:

We can round this off to 3d2

If the pump were running at 2900 rpm you would get



How do we use this information? You can combine this formula with your pressure to head and come up with an estimate to

see if an operating pump is operating close to its BEP(best efficiency point ). As

ure gage reads 20 psi. The pump is pumping the difference between these readings, so the pump is pumping 100 psi.

he pressure to head conversion is:

knowledge of how to convert

an example:

In the inch system a pump discharge pressure gage reads 120 psi. The pump suction press

At its BEP(best efficiency point) the pump should be running between 80% and 85% of its shut off head. 100 psi is 83% of 120 psi. T

The pump has an 8.5 inch impeller running at 3500 rpm. The shutoff head would be (8.5 inches)2 x 4 = 288 feet. Pretty close!

In the metric system we can make the calculation for a 295 millimeter impeller

The pump discharge pressure gage reads 10 bar The pump suction pressure

ing 9 bar

turning at 2900 rpm

gage reads 1 bar The pump is pumping the difference between these readings so the pump is pump

At its BEP(best efficiency point) the pump should be running between 80% and 85% of its shutoff head. 9 bar is 83% of 10.8 bar. The pressure to head conversion is:

106 meters shut off head. The pump has a 295 mm impeller running at 2900 rpm. The shutoff head would be (2.95 decimeter)2 x12 =104.4. Pretty close!

Rules of thumb for pumps

ok at the manufacturer's published pump curve. The problem is that you do not always

mp companies test their pump to determine its performance, they have no need for general guide lines or "rules of thumb."

PUMP BASICS

p (inch sizes) o At 1750 rpm. Shut off head = Diameter of the impeller squared

3500 rpm. Shut off head = Diameter of the impeller squared x 4 o For other speeds you can use the formula : Shut Off Head = D2 x (new

ber. (6,25)

If you want to know a pumps capabilities the rules are simple, lo

have the curve available. Pu

Over the years I have accumulated many of these rules to help me estimate pump performance, here are a few of them:

• How to estimate the shut off head of a pum

o At

rpm / 1750)2 • Estimating metric head is a little bit more involved, but it still works:

o Measure the shaft in mm. ( as an example: 250 mm ) o Mark off two places. (2,5) o Square the numo For 1450 rpm, multiply by 3 (18,75) o Add 10 % for the answer in meters. (21 meters )

40

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would multiply by 12 instead of 3. Although se formulas you cannot estimate the

r that.

nter of the inboard bearing to the center

• modulus of elasticity changing shaft

t solution. When pump

e speed of a pump you will get twice the capacity, four times the

ft has only a small portion of the conductivity of a carbon

ack to the bearing oil.

curves for head, capacity, and

6% stock

r less e settings are determined by the pump manufacturer and

0,2 to 0,5 mm) You lose 1% of the 0,05 mm) you miss this setting.

wear rings less than two inches (50 mm.) in outside diameter.

o NOTE: For 3000 rpm, youyou can estimate shutoff head with thepump capacity. You will need the pump curves fo

• The pumps best efficiency point (B.E.P.) is between 80% and 85% of the shut off head. At this point there is little to no radial thrust on the impeller. Also the "power in" is closest to the "power out".

• The L3/D4 ratio should be below 60 (2.0 in metric) to prevent excessive shaft bending. To calculate it for end suction centrifugal pumps :

o L = length of the shaft from the ceof the impeller (inches or millimeters). Caution: do not use centimeters, the numbers will come out wrong.

o D = diameter of the shaft (under the sleeve) in the stuffing box area (inches or millimeters) Do not use centimeters.

Since most shaft materials have a similar materials will not prevent shaft bending when you operate off of the B.E.P. Lowering the L3/D4 is the only logical and efficienmanufacturers discuss operating off of the B.E.P. they relate problems to the heat that will build up in a minimum flow condition and ignore the problems with shaft bending.

• A double suction pump can run with 27% less N.P.S.H. or at a 40% faster speed without cavitating.

• If you double thhead and it will take eight times the horsepower to do it.

• A stainless steel shasteel shaft. This is very important when you are pumping at elevated temperatures because we do not want to transmit the high temperature b

• If you double the speed of a pump you will get almost four times the shaft whip, wobble or run out and eight times the wear.

• Multistage pumps reduce efficiency 2% to 4%. • In many instances an inducer can lower Net Positive Suction Head Required by as

much as 50% . • If you are pumping paper stock, modify the

efficiency as follows: o 0.725 foro 0.825 for 5.5% stock o 0.90 for 5% o 0.94 for 4.5% o 0.98 for 4% o 1.0 for 3.5% o

• Open impeller clearancnormally run between 0.008" and 0.015" (pumps capacity for each 0.002" (

• Wear ring clearances are very similar to impeller clearances, but you lose 1% pump capacity for each 0.001" (0,025 mm) of wear. A typical clearance would be 0.003 inch/inch diameter with 0.010 inches (0,3 mm) minimum clearance for

42

face seals in these

tive face

e bearing race of a properly installed

F) temperature increase. This corresponds to :

Th il is not being contaminated by wa

• Packing leakage

earing case.

running at 1750 rpm. would cover about 100,000 miles ning

tomatic transmission oil every 25,000 miles ( 40,000 kilometers)

umping temperature exceeds 200 degrees Fahrenheit (100° C). This design will allow the wet end of the pump to expand in

ions instead of from the feet up, destroying the wear rings.. • Try to buy pumps with a Suction Specific Speed (SSS) below 8,500 (10,000

• Bearing grease or lip seals have a design life of less than 2000 hours. In a constantly running pump this would be only 83 days. These seals will also damage the expensive shaft and place a stress point at the maximum bending moment arm. Substitute non fretting labyrinth seals, or positivelocations. It is a good idea to install them in electric motors also to prevent moisture from entering and damaging the motor windings and bearings.

• Do not use a vent on the top of the bearing case. At shut down the outside moisture will enter the bearing housing through this vent. Let the moisture attempt to enter the case through the labyrinth seals instead, they will do a better job of directing the moisture to the external drain hole. If you install posiseals you can forget about this problem.

• The axial clearance in a bearing is ten times the radial clearance. This is the reason proper installation is so critical. If the bearing is over compressed the bearing balls will distort and roll instead of spin causing excessive heat and premature failure. The temperature at thbearing is at least 10 degrees Fahrenheit (5° C) higher than the oil sump temperature.

• The life of bearing oil is directly related to its temperature. The rule of thumb used by the SKF Bearing Company is that the service life of an oil is specified as 30 years at 30 degrees Centigrade (86° F) and is cut in half for each 10 degree Centigrade (10

o A life of 3 months at 100 C. (212 F.) o A life of 6 months at 90 C. (195 F.) o A life of 12 months at 80 C. ( 176 F.)

ese numbers assume that the lubricating oter from one or all of the following sources:

• The water hose used to wash the packing leakage away from the pump area. • Aspiration, as moisture laden air enters the b

An automobile engine (160,000 kilometers) every 2000 hours (83 days in the life of a constantly runpump ). Auto manufacturers recommend changing their au

APPLICATION

• Use Centerline pump designs when the p

two direct

metric) Do not buy pumps with a SSS over 12,000 ( metric 16,500) unless you are pumping hot water or mixed hydrocarbons. If you have a double suction pump you can divide the SSS number by 2

43

. If you have a high static or pressure

e best efficiency point.

ly .

d on long shaft vertical pumps to prevent

ad greater than 650 feet (198 meters) and more

ency vibrations and low frequency vibrations at reduced flow

hould be at least 10 diameters of pipe between the suction of the pump and the first elbow. This is especially critical in double ended pump designs as the

inlet flow can cause shaft thrusting, and subsequent bearing problems. • Substituting a globe valve for a gate valve in a piping system is similar to adding

in the

speed cuts in. You will

side down.

• Do not specify a pump with the largest impeller available . Give yourself an additional 5% or 10% you might need it.

• The maximum viscosity a centrifugal pump can handle would be a product similar to 30 weight oil at room temperature.

• Use a variable speed pump if your head is mainly system head. Circulating hot or cold water would be typical applicationshead, as is the case with a boiler feed pump, the variable speed will not be of much help in keeping you on or near th

• Pumps piped in series must have the same capacity (impeller width and speed) • Pumps piped in parallel must have the same head (impeller diameter and speed ) • Use a rotary positive displacement pump if your capacity is going to be less than

20 gpm.(4,5 cubic meters per hour) • A centrifugal pump can handle 0.5% air by volume. At 6% it will probab

become air bound and stop pumping. Cavitation can occur with any amount of air• Use double volute pumps any time your impeller diameter is 14 inches (355 mm)

or greater. They should also be useexcessive shaft movement that will cause problems with the packing, seals, bearings and critical dimensions.

• A Vortex pump is 10% to 15% less efficient than a comparable size end suction centrifugal pump.

• The A.P.I. (American Petroleum Institute). sixth edition states : High energy pumps, defined as pumping to a hethan 300 horsepower (224 KW) per stage, require special consideration to avoid blade passing frequrates.

PIPING ETC..

• There s

turbulent

another 100 feet (31 meters) of piping to the system. On the discharge side of the pump this will cause the pump to run off of its B.E.P. with a resultant shaft bending. On the suction side of the pump it will probably cause Cavitation.

• After the pump and motor have been aligned, dowel both the pump and the motor to the base plate. Be sure to dowel only the feet closest to the coupling, allowing the outboard ends to expand with temperature changes.

• Check impeller rotation after installing the pump. Do not assume it will turn correct direction. I have heard about two speed pumps with the second speed wired backwards. They will drive you crazy because the pump will often meet its head requirement but not the capacity when the secondalso notice excessive noise at this time.

• Use eccentric reducers rather than concentric reducers at the pump suction. Concentric reducers will trap air. Be sure the eccentric reducer is not installed up

44

d level falling greater than 3 Ft./sec. (1 Meter/ sec.)

t velocities in pipes leaving vessels. Generally greater than 10

apor point. ditions.

onsult the

t be at least one and one half diameters of

TROU

low. o Air is entering at the pump suction.

uid turbulence at the pump suction. • Cavitation damage on the leading edge of the impeller blade indicates internal

below 9000 hat the problem is with the impeller ted when the pump manufacture tried

, as a result, often experience this

of 6 inches (150 mm.) wider.

• Suction piping should be at least one size larger than the suction flange at the pump.

• Vortexing can occur if any of the following conditions are present: o Low liquid levels o Liquio There is a large concentration of dissolved gases in the liquid. o High outle

feet/sec. (3 meters/sec.) o Liquids near their vo High circulation caused by asymmetrical inlet or outlet cono Inlet piping too close to the wall or bottom of the tank. C

Hydraulic Institute Manual or a similar publication for recommended clearances.

o In a mixer, the liquid level musthe blade, above the blade.

BLESHOOTING

• Cavitation damage on the trailing edge of the impeller blade means : o The N.P.S.H. available is too

o There is liq

recirculation. Check the Suction Specific Speed number to see if it is (10,000 metric). Higher numbers mean tshape or adjustment. The problem was creato come up with too low a N.P.S.H. Required.

• Cavitation damage just beyond the cutwater, on the casing and tip of the impeller blade, indicates the impeller blade is too close to the cutwater. This clearance should be at least 4% of the impeller diameter up to a 14 inch (356 mm.) impeller, and 6% greater than 14 inch ( 356 mm.). Some self priming pump manufacturers want a maximum clearance of 1/8" (3 mm) andproblem. A repaired or substituted impeller is often the cause of the problem in a non self priming pump.

• Water in the bearing oil will reduce bearing life 48%. The water enters from packing leakage, wash down hoses, and aspiration caused by the temperature cooling down in the bearing casing after shutdown and moisture laden air entering the bearing case. A 6% water content in the oil will reduce bearing life by as much as 83%

• The mass of the pump concrete foundation must be 5 times the mass of the pump, base plate, and other equipment that is being supported, or vibration will occur.

• Up to 500 horsepower (375 KW), the foundation must be 3 inches (76 mm.) wider than the base plate all around. Above 500 horsepower (375 KW) the foundation should be a minimum

45

sitive face sealing at the bearings, if you could solve the emission problem.

the

ir will enter the

eater drain lication.

the suction

rloading power will occur if the pump is run

Horsepower is measured using the units, foot pounds with one horsepower equal ounds

per minute we are pumping by finding out how much a gallon of our fluid weighs.

the pump is producing and you have foot pounds per minute that can be converted to horsepower.

• Imaginary lines extended downward 30 degrees to either side of a vertical through the pump shaft, should pass through the bottom of the foundation and not the sides.

• The bearing oil level should be at the center of the lowest most ball of a stationary bearing. The preferred choice for bearing lubrication would be an oil mist system with po

• Pipe from the pump suction flange to the pipe rack, not the other way around. • Make sure eccentric reducers are not installed upside down at the pump suction.

The top of the reducer should go straight into the suction flange. • Valve stems, T Branches and elbows should be installed perpendicular to

pump shaft, not at right angles. • Do not use packing in any pump that runs under a vacuum, as a

system through the pump stuffing box.. These applications include : o Pumps that lift liquid. o Pumps that take their suction from a condenser or evaporator. o Any pump that takes its suction from a negative pressure. H

pumps are a typical app• Be sure too vent the stuffing box of a sealed, vertical pump back to

side of the pump or air will become trapped in the stuffing box. The vent must be located above the lapped seal faces.

• If the Specific Gravity of the pumping liquid should increase, due to temperature, there is a danger of overloading the motor and therefore motors having sufficient power should be used. The same ovetoo far to the right of its B.E.P.. This is a very common problem because of the great number of oversized pumps in existence.

How to the calculate the water horsepower coming out of the pump?

to 33,000 foot pounds. Since fluid has weight we can calculate how many p

After you have done that, multiply the gallons per minute you are pumping by 8.33 (the weight of a gallon of water) and then multiply that result by the specific gravity (the weight) of your fluid, and you will have the pounds per minute number you are looking for.

Once you have the pounds per minute you are pumping, you can multiply that number into the feet of head

Please take a look at the following pump curve.

Let's use this chart for our example:

You are using the 13 inch impeller at 1750 rpm and pumping 300 gallons per minute of a fluid with a specific gravity of 1.0 to a head of 168 feet

300 gpm x 8.33 x 1.0 sg. = 2,499 pounds of fluid per minute.

We are pumping this fluid to a head of 168 feet so:

2,499 x 168 = 419,832 foot pounds per minute

Since 33,000 foot pound per minute equals one horsepower. We will divide and get:

419,832 / 33,000 = 12.73 horsepower.

This means that the pump is putting out 12.73 horsepower. Now, the next question is how much actual horsepower is required to do this?

Please take a look at the ascending lines on the bottom of the chart. Each line represents a different size impeller with the top line showing the horsepower required for a 13-inch impeller and the bottom line for a 9-inch impeller. The horsepower required is shown in the left column under bhp. (brake horsepower).

46

Notice that it calls for 20 horsepower to move 300 gallons per minute with a 13-inch impeller.

If the pump were 100% efficient, all you would need would be 12.73 horsepower motor to drive the pump and it would do the job, but motors and pumps are not 100 % efficient because of friction losses and heat generation. This means that our actual efficiency is

12.73 hp out / 20 hp in = 0.64 or 64% efficient

Suppose the specific gravity of the fluid you are pumping is different than 1.0 (cold water). Just plug the new number into the formula and multiply the pump curve bhp by the same number. Using a specific gravity of 1.3, the change would look like this:

300 x 8.33 x 1.3 = 3248.70 pounds per minute

3248.70 x 168 = 545,781.60 foot-pounds per minute

545,781.6 / 33000 = 16.54 water horsepower out of the pump

20 x 1.3 = 26 horsepower is going into the pump

16.54 / 26 = 0.64 efficient

A few things you should know about your pump's piping system

• There should be at least 10 diameters of pipe between the suction of the pump and the first elbow. This is especially critical in double-ended pump designs as the turbulent inlet flow can cause shaft thrusting, and subsequent bearing problems. If an elbow must be installed be sure it is in a plane at right angles to the pump shaft to prevent an uneven flow to both sides of a double suction impeller.


The top of the reducer should go straight into the suction flange. • Piping should be arranged with as few bends as possible. If bends are necessary

use a long radius when ever possible • Valve stems, T Branches and elbows should be installed perpendicular to the

pump shaft, not at right angles. • If an expansion joint is installed in the piping between the pump and the nearest

point of anchor in the piping, It should be noted that a force equal to the area of the expansion joint (which could be a lot larger than the normal piping size) times

47

48

the pressure in the piping will be transmitted to the pump proper. Pipe couplings that do not provide an axially rigid connection have the same affect. If an expansion join or non-rigid coupling must be used, it is recommended that a pipe anchor be installed between it and the pump.

• Be aware that radial forces are being generated in the pump housing from the pressure in the piping system acting on the volute area. The magnitude and direction of the forces is dependent upon the piping arrangement along with the areas and pressures involved.

• It is always a good idea to increase the size of the suction and discharge pipes at the pump nozzle in order to decrease the head loss from pipe friction.


• If increasers are used on the discharge side to increase the size of discharge piping, they should be installed between the check valve and the pump.

• Both a check and gate valve should be installed in the discharge piping with the check valve placed between the pump and the stop valve to protect the pump from reverse flow and excessive back pressure. Manually operated discharge valves that are hard to reach should have some facility for quick closing. A sprocket rim wheel and chain or a remotely operated motor are two alternatives you might consider.

• Suction piping must be kept free of air leaks. • The installation of check valves should be avoided in the suction piping although

they are often used to reduce the number of valves that have to be operated in switching between series and parallel pump operation.

• A foot valve is often installed in the suction piping to aid priming. Do not install them if the pump is operating against a high static head because failure of the driver would allow liquid to rush back suddenly causing water hammer. This is especially true for vertical turbine and submersible pumps that are not designed for use with a foot valve.

• Foot valves should be of the low loss flap type rather than the multiple spring variety and have a clear passage for the liquid at least the same area as the suction piping.

• A horizontal suction line should have a gradual rise or slope to the pump suction. • Cast iron pumps should never be provided with raised face flanges. If steel

suction or discharge piping is used, the pipe flanges should be of the flat face type and not the raised face type. Full-faced gaskets must be used with cast iron flanges.

• The optimum control valve location is within five feet (1,5 meters) of the pump discharge to prevent too much surging of fluid in the system when the discharge is throttled.

• The optimum pipe size will consider the installed cost of the pipe (the cost increases with size) and the pump power requirements (the power required increases with pipe friction)

o Try to limit the friction loss at design flow to 2-5 feet for each 100 feet (1-2 meters for each 30 meters) of pipe).

o To prevent the settling of solids you need a minimum velocity of about 4 to 7 feet per second (1.5 to 2.5 meters per second)

o Velocities of no more than 10 feet (3 meters) per second are recommended in the suction side piping to prevent abrasive wear.

A few rules of thumb for mechanical seals

Before selecting your mechanical seal design there are three things you want to remember:

• All of the seal materials must be chemically compatible with any fluids that will be pumped through the system and that includes solvents, cleaners or steam that might be introduced into the system to flush or clean the lines. It also includes any barrier fluids that are used to circulate between dual mechanical seals.

• The seal faces must stay together. If they open the seal will leak and allow solids to penetrate between the faces where the solids will eventually destroy the lapped surfaces.

• Good seal life is defined as running the mechanical seal until the carbon face is worn away. Any other condition is called a seal failure and is always correctable

The following is offered as a guide when dealing with mechanical seals in general. If possible you should contact the manufacturer for specific recommendations and limits. I have spent the past twenty seven years lecturing about seals and pumps and during that time have picked up a number of rules that are worth remembering. Here are some of the most important:

• Selecting materials - The elastomer ( the rubber part) • There are two temperature limits for a mechanical seal:

o You must not exceed the temperature of the seal components. As an example Ethylene Propylene rubber cannot seal hot fluids in excess of 300° degrees Fahrenheit ( 150° C) without taking a compression set and eventually leaking.

o You must not exceed the temperature limit of the fluid you are pumping. Many fluids will change from a liquid to a gas, solid or crystal at elevated temperature. In almost every case this will cause a seal failure. As an example, petroleum lubricating oil cokes between 250 and 300 degrees Fahrenheit (120° C. to 150° C.) and restricts the movement of the seal components. A Viton® O-ring, in this application would not have been subjected to its temperature limit, but we had the seal failure because we exceeded the temperature limit of petroleum products.

• Halogens will attack Teflon® coated elastomers . Halogens are easily identified because they end in the letters " INE". The list would include Bromine, Chlorine. Astatine, Fluorine, and Iodine. These Halogens will penetrate the Teflon®

49

50

coating and attack the base rubber material causing it to swell and split the Teflon sleeve or coating.

• Most Viton® compounds are attacked by water. Be sure to check if you have the correct one. Remember that steam is another name for water and the steam cleaning of lines is very common in the process industry. Caustic is another common cleaner and caustic contains a high percentage of water also.

• Buna "N" (Nitrile) is an elastomer that has a short shelf life. This is the elastomer that is most often used in Rubber Bellows Seals. The problem is Ozone attack. Ozone is produced by the sparking from electric motors, so it is a very common problem. A typical shelf life for most Buna compounds would be one year.

• If a round O-Ring becomes square in operation (compression set) it is almost always caused by excessive heat. Chemical attack is usually recognized by a swollen and soft elastomer while high heat will produce a shrunken, hard one.

• Chemical attack of the elastomer will usually cause a seal failure within five to ten days. The swollen elastomer will "lock up" the mechanical seal and in some instances, open the lapped seal faces.

Determine the correct O-Ring by one of the following methods:

• Look up the chemical in published O-Ring charts provided by all reputable seal companies. You will find a chart in the chart section of this web site

• Check to see if the plant has any experience with O-Rings, in this fluid, in another seal application. O-Rings can also be found in filters, strainers, valves, flanges, expansion joints etc..

• Test the O-Ring by immersing it into the sealing fluid for one week. If the O-Ring changes weight, shape, or appearance, it is not compatible with the fluid.

• Use a universal O-Ring compound such as Green Tweed's Chemraz, Dupont's Kalrez® or a similar product.

• When choosing an O-Ring, or any other elastomer, be sure to consider any cleaners or solvents that might be flushed through the lines or that could come into contact with the seal. The elastomer must be compatible with these fluids also.

• Never use " glued together" elastomers in a split seal or any "dynamic" application. A hard spot will be created that will interfere with the movement of the dynamic elastomer.

Selecting Materials - The Faces.

• Carbon and most hard face materials have an expansion rate of about one third that of stainless steel.

• Use two hard faces if the product has a tendency to solidify between the seal faces. Never use plated or coated hard faces in these applications. Hard faces are recommended if you find that it is impossible to keep the seal faces together and solids are present in the sealing liquid. Two hard faces are also recommended in the sealing of hydrocarbons that have to pass a "fugitive emissions" test. Coke

51

particles forming between the faces will pull pieces of carbon out of the carbon/graphite face presenting a leak path for fugitive emissions.

• Although many carbon graphite compounds are available unfilled carbons are the best because they are corrosion resistant to almost all chemicals except oxidizing agents and some de ionized water applications. These oxidizing agents will combine with the carbon to form Carbon Monoxide and Carbon Dioxide. The most common oxidizers are oleum, sulfur trioxide, strong bleaches and nitric Acid. You cannot use any form of carbon in these applications. Keep in mind that black elastomers will also be attacked by oxidizing agents because of their carbon content.

• Ceramic vs. ceramic is a good choice for oxidizing chemicals. • If you are going to select plated Tungsten Carbide as a face material, use only the

nickel base Tungsten Carbide. Cobalt base is too hard and can crack with normal seal face differential temperatures. Nickel base, because of its superior corrosion resistance is the preferred material for solid Tungsten Carbide faces also.

• Reaction bonded Silicone Carbide has excellent wear characteristics, but contains up to 17% free silica which can be attacked by many chemicals including caustic. Alpha sintered Silicone Carbide is also available and is Silica free.

• 85% ceramic should never be recommended as a hard seal face as it can break with as little as a 100 degree Fahrenheit (55 C) temperature difference. 99.5% would be a much better choice.

• Plating or coating a seal face will not give it corrosion resistance. Coatings are used for wear resistance and low friction. To get corrosion resistance the outer coating must be at least 1/8" (3 mm) thick. If the base material is not corrosion resistant to the pumping fluid and any cleaners or solvents used in the lines the corrosive will go through the coating and attack the base, causing the plating to come off in sheets.

Selecting Materials - The Metal Parts.

• Be sure to use low expansion metal such as Carpenter 42 or Invar 36 in your metal bellows seal face holder if the product temperature can exceed 400° Fahrenheit (205°C). These low expansion steels will prevent the carbon or hard seal faces from leaking between the face and the metal holder. Needless to say glue or epoxy is not a sensible solution to differential expansion problems.

• If your pump is manufactured from Iron, steel, stainless steel, or bronze, you can probably use a seal manufactured from 316 stainless steel components. The springs or bellows, however, must be manufactured from Hastelloy "C" to avoid problems with Chloride Stress Corrosion.

Sealing Limits

• Use only stationary mechanical seals (the springs do not rotate with the shaft) if the face surface speed exceeds 5000 feet per minute ( 25 M/sec.), but never in a cartridge design unless some method has been provided to insure that the cartridge sleeve is square to the shaft.

52

• Use O-Ring balanced seals in vacuum applications down to 10-2 inches or one millimeter of mercury (1 Torr.). The O-Ring is the only elastomer that can seal both vacuum and pressure. Split seals will work in these applications, but they must be turned around for best operation.

• Any good quality, balanced, O-Ring seal can seal stuffing box pressures to 400 psi (28 bar) and temperatures to 400 degrees Fahrenheit (205° C). There is a compound of Dupont's Kalrez® that is satisfactory to 600 degrees Fahrenheit (370° C), but it is not acceptable at ambient temperatures (it gets too hard).

Application

• A Balanced O-Ring seal will not vaporize the product at the seal face if the stuffing box pressure is at least one atmosphere above the products vapor point.

• The easiest product to seal is a cool, clean, lubricating liquid. All problem chemicals can be placed into several categories. If you know how to seal these categories you should have no trouble making seals work in your applications :

o Products that crystallize (caustic or sugar solutions) o Viscous products (asphalt or molasses) o Products that solidify (polymers or chocolate) o Products that vaporize (hot water or benzene) o Film building liquids (hot petroleum or plating solutions) o High temperature fluids (heat transfer oil or liquid sulfur) o Dangerous products (fire hazard, explosive, radioactive, bacteria) o Non lubricating liquids (solvents or hot water) o Gases and dry running applications (hydrogen) o Dry solids (cake mix or pharmaceuticals) o Corrosive fluids (acids or strong bases) o Cryogenics (liquid nitrogen) o Slurries (river water, sewage, most raw products)

• In addition to these chemical categories there are other sealing problems that include:

o High pressure o Hard vacuum o High speed o Excessive motion

• Dual seals should be balanced in both directions to prevent failure when barrier fluid pressure changes. The practice of using "one direction" seal balance is commonly employed by most seal companies and should be avoided for both safety and reliability.

• Use motion seals on mixers, agitators, sleeve bearing equipment and any rotating device that has motion greater than 0.005" (0,15 mm.) in a radial or axial direction. Pump seals do not work well in these applications because the hard faces are too narrow and the internal seal clearances are too tight.

• Do not use flushing fluid as a coolant in stationary mechanical seals. The coolant will be directed to only one side of the seal and since a stationary seal does not rotate the sliding components the differential temperature can cause the faces to

53

go out of flat. In the case of stationary bellows seals it could cause a bellows rupture.

• The best way to cool a seal is to use the jacketed stuffing box that came as a part of the pump. This jacket will not only cool down the seal area, but will provide the necessary cooling to the shaft so that it will not transmit stuffing box heat back to the bearings.

• The use of steam in a Quench gland is another solution, but not as good as the jacketed stuffing box.

• It is all right to dead end fluid in a stuffing box if a jacketed stuffing box is being used. Do not attempt to recirculate back to the suction side and cool the stuffing box at the same time. When using a jacketed stuffing box it is best to install a carbon bushing in the bottom to act as a thermal barrier the pumping fluid and the seal.

• Do not use rotating, "Back to Back" double seals in dirt or slurry service. The solids will prevent the inner seal from moving forward as the faces wear and if the barrier fluid pressure is lost, solids will penetrate the inner seal faces.

• Be sure to vent vertical pumps back to the suction side of the pump. Air trapped in the stuffing box can cause the seal faces to run hot and in some instances destroy the elastomer.

• Cyclone type separators or "in line filters" are not a good method of cleaning up the fluid in the stuffing box.

• Heat affects a seal several ways: o The faces can be attacked. Plated faces can have the hard coating crack off

and filled carbons can have the binder melted out in high heat. o The elastomer (rubber part) has a temperature limit determined by the

compound used. o The corrosion rate of all liquids increases with temperature. o Thermal expansion can cause seal face loads to alter and seal face flatness

to change. o Many products will change from a liquid to a solid or gas in the presence

of high temperature. If this should occur between the seal faces, they can be blown open.

• Do not be tempted to put the mechanical seal outside of the stuffing box to keep the springs out of the fluid. As the face wears the seal must move into the slurry where it will eventually "hang up" and leak. In these applications centrifugal force is throwing solids into the lapped faces and if there is excessive pressure in the system the seal faces will be blown open.

• When choosing the pressure range of a mechanical seal be sure to consider the stuffing box pressure not the pump discharge pressure. Very few seals will ever see discharge pressure.

Technical

• Seals lapped to less than three helium light bands ( 0.000034") inches or 1,0 microns) should not show visible leakage. Visible leakage occurs at about 5 light bands.

54

• A typical mechanical seal face load would be 30 psi. (0,2 N/mm2) when the carbon is new and 10 psi. (0,07 N/mm2) when the carbon is fully worn away. You must never guess as to how much to compress a mechanical seal. Either take the information from the seal print or calculate the correct length from the above information.

• Both rotating and stationary metal bellows seals require vibration damping. Elastomer seals do not experience this vibration problem because the elastomer touching the shaft is a natural vibration damper. Vibration can be either harmonic or caused by poor lubricating fluids (slip stick)

• Use only non fretting seal designs. Shafts and sleeves cost too much to ignore this severe problem.

• Carbon throat bushings should have a shaft clearance of 0.002 inches/inch (0,002 mm/ millimeter) of shaft diameter. If they are to be used as a support bearing you should cut the clearance down to 0.001 inches/ inch (0,001 mm/millimeter) of shaft diameter.

• It is not necessary to lubricate seal faces at installation. If the product you are sealing can vaporize between the faces and cause freezing then you must remove any lubricant that might have been placed there by the manufacturer.

• Balanced mechanical seals consume about one sixth the horsepower of packing. Packing a pump would be like running your automobile with the emergency brake engaged. The car would run, but the fuel consumption would be high.

• Single spring seals are wound in either a right or left handed direction. Check to see if your seal has a problem in keeping the faces together because of the spring winding.

• Open impeller pumps require impeller adjustment. Use only cartridge or split seals in these applications. Do not use seals that locate against a shoulder or set screw to the shaft, as the face load will change when the impeller is adjusted.

• Do not relap the carbon face unless it is an emergency. Seal face opening is a common seal failure. When the faces open solid particles imbed them selves into the carbon face and will be driven in even further during the lapping process. If you must relap in an emergency never use lapping powder, as the abrasive particles will imbed into the soft carbon.

• You cannot balance an inside seal by removing material from the carbon face. To get seal balance you must do one of the following:

o Use a stepped sleeve with rotating seals. o Let the carbon slide in a case that is sealed to the shaft. o Use a metal bellows. The balance is not perfect, but good enough. o Use a stationary seal design, they require no stepped sleeves.

• Seal face hardness is a confusing subject because of the various measuring scales employed. The two most common are Rockwell "C" and Brinnell. If you divide the Brinnell scale by ten (10) it is almost equal to the Rockwell "C" scale.

• Avoid oil as a barrier or buffer fluid between two mechanical seals. Most petroleum base and other oils have a low specific heat (0.2 - 0.4) and combined with poor conductivity (0.5 of water) makes them a poor choice compared to fresh water. If oil is mandatory, a clean heat transfer oil would be your best choice.

• A convection tank can often be used between two balanced O-Ring seals. If you use unbalanced seals the heat generated by this type of seal is usually excessive for convection cooling. Contact the seal manufacturer for his recommendations concerning speed, diameter, face combination and pressure limits for convection cooling. If convection is not satisfactory, a pumping ring or forced lubrication is another option.

• If you decide to repair your mechanical seals in house, be sure to purchase the parts from the original manufacturer. If you decide to have them repaired send them back to the original manufacturer. It is important that the seal be rebuilt with the original materials and it must meet the original tolerances. This information is not available from the manufacturer because of product liability problems.

• O-ring seal designs can tolerate three to four times the "run out" capability of sliding or pusher seals incorporating wedges, chevrons, U- cups etc..

• Oil on the seal faces can cause the faces to stick together during long periods of non running. If you do not intend to run the equipment soon remove any oil that might be on the seal faces during the assembly procedure.

A quick reference to prevent potential seal and pump problems:

The biggest advantage of experience is you have hopefully learned what can get you into trouble. The following information has been explained in detail in previous technical papers, but I still see the same problems re-occurring on a daily basis.

Take a few minutes and look at the following. It might save you a seal or pump failure.

MATERIALS

• Carbon seal face. Any form of carbon is usually not acceptable in the following applications:

o Oxidizers, they combine with carbon to form CO & CO2 o Halogens (most of them end in the letters "ine") chlorine, bromine,

fluorine, astintine & iodine o Where color contamination can be a problem. o Some de-ionized water applications. o Hot petroleum products if you are concerned about fugitive emissions.

• A special carbon is used for cryogenic and hot dry air applications. Moisture is needed to make the graphite release from the carbon-graphite mixture, and in these applications the needed moisture is not present. A special carbon with an imbedded organic is made to satisfy these applications.

• Ceramic grade 99.5 is not a satisfactory hard face in hot applications because of its poor thermal conductivity. Alpha grade silicone carbide or tungsten carbide are much better choices.

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• Ethylene Propylene Rubber O-Rings will be attacked by petroleum products and this includes any petroleum grease that might be put on the O-ring during the installation process.

• Kalrez® grade 3018 is not satisfactory if the temperature is below 600°F. (315°C.) The material is too hard at these lower temperatures.

• Nickel base tungsten carbide can cause galvanic corrosion problems with stainless steel shafts.

• Reaction bonded silicone carbide is not satisfactory for caustic or most high PH materials.

• Viton® O-rings are not generally satisfactory in water based fluids. This also includes steam cleaning or flushing the lines with water based caustic solutions. Grade 747-75 fluorocarbon is O.K. if the water is cold, but ethylene propylene rubber is still your best choice as long as the temperature does not exceed 300°F (150°C.).

• White Chemraz is not recommended for most high PH fluids. Do not use it with: o Acetaldehyde, Ammonia + Lithium metal solution, Aqua Regia, Black

liquor, DI water, Ethyl Formate, Ethylene Oxide, FC 75, Freon 113 -114 - 114B2 - 115 - 142B- C318 - PCA - TF, Fuming Sulfuric Acid, Green Sulfate Liquor, KEL-F- Liquids, Lye, Magnesium Hydroxide, Red Fuming Nitric Acid, Potassium Hydroxide, Sodium Hydroxide (Caustic), Fuming Sulfuric Acid, and White Liquor.

• If you choose the wrong elastomer it will be attacked by the fluid and break down. For the first few days the seal will work very well because the elastomer has become "slimy" and moves easily. The elastomer will then "swell-up" and lock-up the moveable seal components.

APPLICATION

Remember that chemical attack can be accelerated by temperature, fluid concentration, and stress. Past plant experience is your best indicator of what seal and pump materials to use.

• Ammonia compressor; use Neoprene for the O-ring because the fluid is a combination of ammonia and petroleum oil.

• Black Liquor, as found in paper mill applications can be either sulfite or sulfate. Sulfate (high PH) is the most common and ethylene propylene can be used for the O-ring material if the temperature is below 300°F (150°C). If the temperature is too high, Kalrez is a good choice. White Chemraz is not recommended in these higher temperature caustic applications.

• Boiler feed pump applications vary a great deal. In some cases they are nothing more than a simple hot water application, but in other instances a very high pressure is involved. In any case, cooling is needed in the stuffing box to insure long seal life. High pressure applications also require a heavy duty seal design.

• Caustic. If the concentration is over 50% Monel metal will probably be needed. The metal selection depends upon the temperature and stress.

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• Ethylene Oxide will penetrate into most elastomers and explode out the other side of the O-ring. Use two seals and pressurize between them. Ethylene Oxide is a dangerous product, so two seals should be used in any case.

• Halogens attack most carbon faces and will penetrate the Teflon® encapsulated O-rings like Vanway, Creavey and & 76 style.

• Hot oils. Coking is always the problem. The seal area must be cooled. Coking is a function of temperature and time and is independent of the presence of oxygen. If you want to seal fugitive emissions you will have to go to two hard faces. Even the best of carbons show some blistering in these applications. In other words, a metal bellows seal will not eliminate the need for stuffing box cooling.

• High temperature applications. Most metal bellows seal designs incorporate a low expansion holder (Invar 36 or Carpenter 42) to retain the carbon face. This holder is also frequently used as a vibration damper to prevent seal face separation problems caused by "slip stick." If you lose cooling in these applications the pump shaft expands at a rate three times that of the low expansion steel vibration damper and can cause the seal faces to be pulled open.

• Kaoline (china clay) will penetrate lapped seal faces because the solids are less than one micron in size. You will need two seals with a pressurized barrier fluid between the seals. Water is a good choice for this barrier fluid.

• Latex balls up between the seal faces. Dual seals with a pressurized water barrier fluid have been used in this application, and non contacting gas seal seem to be the current choice, but flushing with a small amount of cold water seems to be the only satisfactory solution to this application.

• Paper stock always requires a small amount of flushing water. You cannot use suction recirculation and centrifugal force to separate the stock from the water because of the stock's low specific gravity. If the pump is trying to "lift" paper stock it will almost always cavitate.

• Pipe line applications almost always involve high pressure. Heavy duty seals should be used in these applications.

• Products that freeze (cryogenic). Watch out for moisture outboard of the seal. Dual seals with anti-freeze circulating in a convection tank is your best bet. Do not put any grease on the seal faces. It will freeze also.

• Salt water. Coat the O-rings and all clamped surfaces with Zinc Oxide paste to prevent corrosion at these locations.

• Sulfuric acid. Alloy 20 metal is usually needed for these applications. Any leakage will cause severe corrosion as the product is diluted.

CONVERTING FROM PACKING TO SEALS

Horizontally split pumps:

• Suction recirculation will not work if the stuffing box is at suction pressure. Most single stage designs fit into this category

• The face of the stuffing box must be resurfaced to get a good gasket seal.

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• If you are making a new gasket between the casing halves, be sure to have it extend outside the stuffing box face and then trim it flush after the halves are tightened together.

• Be sure to seal between the sleeve and the impeller. This is a potential leak path after a mechanical seal is installed.

• Some sleeves terminate under the seal. Check that you will not have a corrosion problem if the sleeve and shaft are different materials,

• Sometimes a new gasket will extrude into the sides of the stuffing box when the two halves of the pump are bolted together. The gasket can then rub against the side of the seal interfering with its movement.

• You will need either a stationary mechanical seal or some type of self aligning feature to seal these pumps successfully.

Flyte sewage pumps can be converted to a single mechanical seal if a special adapter is made. It's worth the problem. You only have to seal the bearing cavity in this application

MISCELLANEOUS

• Discharge recirculation can act as a sand blaster against the seal body. This can be a big problem with the thin metal plates found in metal bellows seals.

• Dual seal barrier or buffer fluid. Oils should be your last choice as a barrier or buffer fluid because of oils' low specific heat and poor thermal conductivity. You will definitely need a pumping ring if you are going to use a convection tank.

• Quenching. An excess of water or steam can easily get into, and ruin the bearings. • Suction recirculation is not affective in the following:

o Duriron pumps, because of their semi- open impeller design. o If the fluid is close to its vapor point. flashing will occur when the stuffing

box pressure drops. o If the specific gravity of the solids is lower than the fluid. If the solids

float, centrifugal force will throw the liquid to the outside leaving the solids against the seal components. Paper stock is a good example of this.

o Single stage, double ended pumps where the stuffing boxes are at suction pressure.

• Seal set screws are normally manufactured from corrosion resistant materials and are therefore softer than normal set screws. This means they can slip if reused. You can substitute hardened set screws in most cartridge seal applications.

• Do not use any type of set screw on non-metallic shafts. Seals must be clamped to the non metallic shaft or sleeve.

• Split seal designs. Most can seal either a pressure or a vacuum, but not if the application alternates between them. You can run into this problem in some mixer applications.

• Troubleshooting hints o Are other seals working in this application? If they are, you know the

materials are alright. Now you must decide what is different about this application.

o Has the seal been repaired? You may be looking at a rub mark, discoloration, or corrosion that is not relevant to this application.

CENTRIFUGAL PUMPS

• Do not let the welder use the pump as an electrical ground. You can ruin the seal or bearings in the process.

• Pumping off of the best efficiency point will not excessively deflect the shaft with the following centrifugal pump designs:

o Double volute casings. o Multi stage designs. o Diffuser or turbine pump designs.

• Be sure to level the pump when you do an alignment. • If you trim the impeller, file the tips and re balance the assembly. • The next time that you look at the pump discharge gauge, remember that the

pump pumps the difference between the suction and discharge heads. You must subtract a positive suction head to determine what head the pump is really creating.

• Bearing lip or grease seals have a useful life of less than 90 days and will cut and score the shaft because of fretting corrosion.

• Never cool a bearing housing because it will shrink and over compress the bearing. Cool only the bearing oil.

• Flushing the system with steam or a cleaner seldom flushes out the stuffing box of the pump.

• Do not circulate shop water through the cooling jacket on a high temperature pump. Condensate or low pressure steam is a better choice. Be sure to install a thermal bushing in the end of the stuffing box to get effective temperature control in the seal area. Make sure you come into the bottom of the jacket and out the top to vent any air that might be trapped in the jacket.

What is the best pump and seal technology?

The "Best Technology" phrase comes up in recent government regulations and every day plant conversations. So what is the best Mechanical Seal and Pump Technology available today? Here is my opinion:

SEAL TECHNOLOGY

Materials

• Identifiable face materials compatible with the fluid to be sealed and any cleaners or solvents put through the lines.

• Materials able to handle the full temperature range of the product you are sealing. • Viton® compatible with water.

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• Hard faces that are not sensitive to temperate change or caustic cleaners. • Unfilled carbon graphite seal faces • No elastomers with shelf life. • No stainless steel springs or bellows.

Design

• The seal should shut with spring and system hydraulic pressure. • Hydraulically balanced designs for low heat generation. • Two way balance in dual seal designs. • Built in pumping ring for cartridge dual seals. • Tandem configuration in dual seal designs. No rotating "back to back" designs. • Stationary configuration for non-cartridge applications. • Self aligning design for stationary cartridge versions. • Springs designed out of the fluid. • The elastomer should move to a clean surface as the faces wear. • No spring loaded elastomers. • Non fretting designs. • Independent of shaft tolerance and finish • Static elastomer located away from the seal face • Cartridge sleeve sealed at wet end. • Vibration damping of the seal face. • Seal should be located close to bearing support. • No elastomer in the seal face. • Faces in compression. • Wide operating range • Low hysteresis. • Equal & opposite clamping of stationary face. • Sealing fluid located at the outside diameter of the seal faces • Leak detection capability • Independent of shaft finish and tolerance • Compensate for thermal expansion and adjustments. • Meet fugitive emission standards. • Simple installation. • Eliminate all elastomers if possible • Short length leaving room for a shaft support bushing. • Finite element analysis of all components. • A method of supporting the shaft in the event of a bearing failure. • Trapped gaskets.

OTHER

• Packaging to survive a one meter drop. • Back up sealing. • Built in seal face vent for vertical applications. • No glued elastomers in split seal configurations.

BEST PUMP TECHNOLOGY

• Low shaft diameter to length ratio (less than 60 L3 /D4) . • Large operating window • C or D frame adapter to simplify driver to pump alignment • Centerline design for thermal expansion. • Oversize stuffing box. • Adequate bearing retention (no snap rings). • Positive bearing sealing. • Oil level indication. • Oil cooling availability. • Low NPSH. • Double volute to prevent shaft deflection. • Suction specific speed number below 8500. • Dynamically balanced rotating assembly. • Impeller specific speed number selected for the application. • Duplex metal impeller. • Impeller investment cast. • Adjust impeller from the wet end to prevent seal face load change.

A new method of troubleshooting centrifugal pumps and mechanical seals:

One of the U. S. based Japanese automobile manufacturers has a unique method of troubleshooting any type of mechanical failure. The system is called the "Five Whys". It is a simple but powerful concept, nothing has been solved until the question "why ?" has been asked at least five times and a sensible answer has been given for each of the "why" questions. As an example:

1. Why did the seal fail?

• The lapped faces opened and solids penetrated between them. (solids can't get in until the faces open)

2. Why did the faces open?

• The set screws holding the rotary unit slipped due to a combination of vibration and system pressure.

3. Set screws are not supposed to slip. Why did the set screws slip?

• The seal was installed on a hardened sleeve.

4. Why was the seal installed on a hardened sleeve?

• This was a packing conversion and a stock sleeve was used.

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5. Why couldn't the mechanic tell the difference between a hardened sleeve and a soft one?

• They were both stored in the same bin.

6 Why were they stored in the same bin?

• Because they had the same part number.

7. Why did they have the same part number?

• They should have had different part numbers. Once that problem is corrected, the failures will stop.

Now you get the idea! Needless to say you may have to go further than just five "whys". Let's try another example:


• The pump was cavitating and the vibration caused the carbon face to crack.

2. Why was the pump cavitating?

• It did not have enough suction head.

3. Why didn't it have enough suction head?

• The level in the tank got too low.

4. Why did the level in the tank get too low?

• I don't know.

You have not finished five "whys" so you better go find out why the level in the tank go too low or the problem is going to repeat its self. In the above example the float got stuck on a corroded rod, giving an incorrect level indication.

One more example should do it:

1. Why did the seal start to leak?

• The elastomer got hard and cracked.

2. Why did the elastomer get hard and crack?

• It got too hot.

3. Why did it get too hot?

• The pump stuffing box ran dry.

4. Why did the stuffing box run dry?

• It was running under a vacuum and it was not supposed to.

5. Why was it running under a vacuum?

• A Goulds pump impeller was adjusted backwards to the back plate and the impeller pump-out rings emptied the stuffing box.

6. Why was it adjusted backwards?

• Most of the pumps in the facility are of the Duriron brand and they normally adjust to the back plate. The mechanic confused the impeller adjustment method. He has since been retrained

This is a powerful trouble shooting technique. I hope you make good use of it.

Operation practices that cause frequent seal and bearing maintenance problems

Wouldn't it be wonderful if the plant operation and maintenance departments could work independently? The fact of the matter is that there are three types of problems we encounter with centrifugal pumps and poor operation is one of them. If you are curious, the other two are design problems and poor maintenance practices.

Seals and bearings account for over eighty five percent (85%) of premature centrifugal pump failure. In the following paragraphs we will be looking at only those operation practices that can, and will cause premature seal and bearing failure. Design and maintenance practices will be discussed in other papers in this series.

When pumps were supplied with jam packing, the soft packing stabilized the shaft to prevent too much deflection. In an effort to save flushing water and to conserve power, many of these same pumps have since been converted to a mechanical seal and the radial stabilization the packing provided has been lost.

The bad operating practices include:

Running the pump dry will cause over-heating and excessive vibration problems that will shorten seal life. Here are some of the common reasons why a pump is run dry:

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• Failing to vent the pump prior to start-up. • Running the tank dry at the end of the operation cycle. • Emptying the tank for steaming or introduction of the next product. • Running on the steam that is being used to flush the tank. • Starting the standby pump without venting it. Venting a hazardous product

can cause a lot of problems with the liquid disposal. Many operators have stopped venting for that reason.

• Tank vents sometimes freeze during cold weather. This will cause a vacuum in the suction tank, and in some cases could collapse the tank.

• Sump fluids are often dirty, corrosive or both. The control rods for the float switch will often "gum up" or corrode and give a false reading to the operator. He may think that there is an adequate level, when in fact, the tank is empty.

Dead heading the pump can cause severe shaft deflection as the pump moves off of its best efficiency point (B.E.P.). This translates to excessive heat that will affect both the seal and the bearings as well as causing the seal faces to open, and the possibility of the impeller contacting the volute when the shaft deflects.

• Starting the centrifugal pump with a shut discharge valve is standard practice with many operation departments. The concern is to save power without realizing the damage that is being done to the mechanical seal, impeller, wear rings and bearings.

• Some pumps are equipped with a recirculation valve that must be opened to lessen the problem, but many times the valve is not opened, or the bypass line is clogged or not of the correct diameter to prevent the excessive head. Another point to remember is that if the bypass line is discharged to the suction side of the pump the increased temperature can cause cavitation.

• After a system has been blocked out the pump is started with one or more valves not opened.

• Discharge valves are shut before the pump has been stopped.

Operating off of the best efficiency point (B.E.P.). Changing the flow rate of the liquid causes shaft deflection that can fail the mechanical seal and over-load the bearings.

• Starting the pump with the discharge valve closed to save power. • The level in the suction tank is changing. Remember that the pump pumps

the difference between the discharge and suction heads. If the suction head varies, the pump moves to a different point on its curve.

• Any upset in the system such as closing, throttling or opening a valve will cause the pump to move to a new point on the curve as the tank fills.

• Pumping to the bottom of a tank will cause the pump to move to a different point on the curve as the tank fills. Some systems were designed for a low capacity positive displacement pump and have since been converted to a

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centrifugal design because of a need for higher capacity. Centrifugal pumps must discharge to the top of the tank to prevent this problem.

• If the discharge piping is restricted because of product build up on the inside walls, the pump will run throttled. This is one of the reasons that it is important to take periodic flow and amperage readings.

• Increasing the flow will often cause cavitation problems.

Seal environmental controls are necessary to insure long mechanical seal life. It is important that operations understand their function and need because many times we find the controls installed, but not functioning.

• Cooling-heating jackets should show a differential temperature between the inlet and outlet lines. If the jacket clogs up, this differential will be lost and seal failure will shortly follow.

• Barrier fluid is circulated between two mechanical seals. There may or may not be a differential temperature depending upon the flow rate. If a convection tank is installed, there should be a temperature differential between the inlet and outlet lines. The line coming out of the top of the seal to the side of the tank should be warmer than the line from the bottom of the tank to the bottom of the seals, otherwise the system is running backwards and may fail completely. The level in the tank is also critical. It should be above the tank inlet line or no convection will occur. Some convection tanks are pressurized with a gas of some type. Many original equipment (O.E.M.) seal designs will fail if this differential pressure is lost.

• Some seal glands (A.P.I. type) are equipped with a quench connection that looks like the seal is leaking water or steam. If there is too much steam pressure on this quench connection, the excessive leakage will get into the bearings causing premature failure. The steam is often used to keep the product warm to prevent it from solidifying, crystallizing, getting too viscous, building a film on the faces etc. Operating people frequently shut off the quench to stop the condensate from leaking.

• Flushing fluids are used for a variety of purposes, but most of the time they are used to get rid of unwanted solids. The flush can be closely controlled with a flow meter or throttling valve. The amount of flush is determined by the seal design. As an example, those designs that have springs in the product require more flush.

• It is important to check that the stuffing box has been vented in vertical pumps. The vent should be coming out of the seal gland and not the stuffing box lantern ring connection.

There are some additional things that all operators should know to insure longer rotating equipment life. As an example :

• Mechanical seals have an 85% or more failure rate that is normally correctable. This is causing unnecessary down time and excessive operating expense. Seals should run until the sacrificial carbon face is

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• There are five different causes of cavitation. • You should know where the best efficiency point (B.E.P.) is on a particular

pump, and how far it is safe to operate off the B.E.P. with a mechanical seal installed.

• You should be aware that washing down the pump area with a water hose will cause premature bearing failure when the water penetrates the bearing case.

• Learn about the affect of shaft L3/D4 on pump operation. • Know how the pumped product affects the life of the mechanical seal and

why environmental controls are necessary. • If you are not using cartridge seals, adjusting the open impeller for

efficiency will shorten the seal life. In most cases the seal will open as the impeller is being adjusted to the volute. Durco pumps are the best example of the exception to this rule. The popular Durco pumps adjust to the back plate causing a compression of the seal faces that can create mechanical seal "over heating" problems.

• Cycling pumps for test will often cause a mechanical seal failure unless an environmental control has been installed to prevent the failure.

• Mechanical seals should be positioned after the impeller has been adjusted for thermal growth. This is important on any pump that is operated above 200°F (100°C) or you will experience premature seal failure.

• Some elastomers will be affected by steaming the system. A great deal of caution must be exercised if a flushing fluid such as caustic is going to be circulated through the lines or used to clean a tank. Both the elastomer and some seal faces (reaction bonded silicone carbide is a good example) can be damaged. If the elastomer is attacked, the failure usually occurs within one week of the cleaning procedure.

• The stuffing box must be vented on all vertical centrifugal pumps or otherwise air will be trapped at the seal faces that can cause premature failure of many seal designs.

• Most original equipment seal designs cause shaft damage (fretting) necessitating the use of shaft sleeves that weaken the shaft and restrict pump operation to a narrow range at the B.E.P..

Here are a few common misconceptions that cause friction between maintenance and operation departments

• Shutting the pump discharge valve suddenly, will blow the seal open. • All ceramics cold shock. • High head, low capacity consumes a lot of power. • The pump must come into the shop to change a mechanical seal. • If you use two hard faces or dual mechanical seals in slurry applications,

you will not need flushing water with its corresponding product dilution.

• If you use metal bellows seals for hot oil applications, you will not need the stuffing box cooling jacket operating.

• It is O.K. to use an oversized impeller because throttling back will save power.

A few more thoughts on the subject

• Operators should receive proper schooling on the trouble shooting and maintenance of pumps. In the military and many modern plants, the operator and the maintenance mechanic are often the same person. If the operator knows how the pump works he will have no trouble figuring out the solution to his problem. Too often he is told to keep the flow gage at a certain point, or between two values without understanding what is actually happening with the equipment. If the operator recognizes cavitation he can tell the maintenance department and help them with their trouble shooting.

• As you wander around the plant look out for painters that paint the springs of outside and double mechanical seals. There is a trend to putting two seals in a pump for environmental reasons and the painting of springs is becoming a common problem.

• If someone is adjusting the impeller make sure he is resetting the seal spring tension at the same time.

• If the pump is getting hot or making excessive noises, report it immediately. After the failure, it does no good to tell maintenance that it was making noise for two weeks.

• If you are the floor operator it is common knowledge that taking temperature and pressure readings is very boring, especially on those gages that are located in hot or awkward locations. Avoid the temptation to "radio" these readings. From hot to failure is a very short trip.

• Maintenance's favorite expression is "there is never time to do it right, but there is always time to fix it." Try to keep this in mind when the pressure is on to get the equipment running again.

• Do not let cleaning people direct their "wash down" hoses directly at the pump. Water entering the bearings through the lip or grease seals is a major cause of premature bearing failure. Most water wash downs are used to dilute and wash away seal leakage. Stop the leak and you have eliminated the reason for the hose.

• A great many motor and electrical problems are caused by these same wash down hoses.

• Cooling a bearing outside diameter will cause it to shrink and the bearing will get hotter as the radial load increases. Keep the water hose and all other forms of cooling off of the bearing casing.

The pump is not producing enough head to satisfy the application?

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This is the first paper in a four part series about pump troubleshooting. Let me begin by pointing out that there are a couple of things you must keep in mind when troubleshooting centrifugal pump problems:

• The centrifugal pump always pumps the difference between the suction and discharge heads. If the suction head increases, the pump head will decrease to meet the system requirements. If the suction head decreases the pump head will increase to meet the system requirements.

• A centrifugal pump always pumps a combination of head and capacity. These two numbers multiplied together must remain a constant. In other words, if the head increases the capacity must decrease. Likewise if the head decreases, the capacity must increase.

• The pump will pump where the pump curve intersects the system curve. • If the pump is not meeting the system curve requirements the problem could be in

the pump, the suction side including the piping and source tank, or somewhere in the discharge system.

• Most pumps are oversized because of safety factors that were added at the time the pump was sized. This means that throttling is a normal condition in most plants, causing the pump to run on the left hand side of its curve.

THE PROBLEM COULD BE IN THE PUMP ITS SELF

• The impeller diameter is too small. o The impeller is running at too slow a speed o You are running an induction motor. Their speed is different than

synchronous motors. It's always slower. The pump curve was created using a variable frequency motor that ran at a constant speed. Put a tachometer on your motor to see its actual speed.

o Your pulley driven pump is running on the wrong pulley diameter. o A variable frequency motor is running at the wrong speed. o Check the speed of the driver if the pump is driven by something other

than an electric motor. • There is something physically wrong with the motor. Check the bearings etc. • Check the voltage of the electric motor. It may be too low. • The impeller is damaged. The damage could be caused by excessive wear, erosion,

corrosion or some type of physical damage. o Physical damage often occurs during the assembly process when the

impeller is driven on or off the shaft with a wooden block and a mallet. Many impeller designs do not have a nut cast into the impeller hub to ease removal.

o Erosion occurs when solids enter the eye of the impeller. The solids can chip off pieces of the ceramic that are passivating the impeller, causing localized corrosion.

o Damage can occur if the impeller to volute, or back plate clearance is too small and the shaft experiences some type of deflection. The original clearance could have diminished with thermal growth of the shaft. Keep in

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mind that some open impellers adjust to the volute (Goulds) while other designs adjust to the back plate (Duriron).

• In an ANSI and similar design centrifugal pumps, the normal thrust towards the volute has bent the snap ring designed for bearing retention. This can allow the rotating impeller to hit the stationary volute.

• Here are some examples of shaft displacement: o Operating the pump too far off the BEP. o Pulley driven applications. o Pipe strain. o Misalignment between the pump and driver. o The shaft could be bent. o The rotating assembly was probably not dynamically balanced.

• The impeller is clogged. This is a major problem with closed impellers. With the exception of finished product, most of what you will be pumping contains entrained solids. Remember also that some products can solidify, or they can crystallize with a change in fluid temperature or pressure.

• Impeller balance holes have been drilled between the eye and the wear rings of a closed impeller. The reverse flow is interfering with the product entering the impeller eye. A discharge recirculation line should have been used in place of the balance holes to reduce the axial thrust.

• The double volute casting is clogged with solids or solids have built up on the surface of the casting.

• The open impeller to volute clearance is too large. 0.017" (0,5 mm) is typical. This excessive clearance will cause internal recirculation problems. A bad installation, thermal growth, or normal impeller wear could be the cause.

o A large impeller to cutwater clearance can cause a problem called discharge recirculation. Wear is a common symptom of this condition.

• If the impeller is positioned too close to the cutwater you could have cavitation problems that will interfere with the head.

• The impeller specific speed number is too high. Lower specific speed numbered impellers are used to build higher heads.

• An impeller inducer was left off at the time of assembly. Inducers are almost always needed with high specific speed impellers. Leaving off the inducer can cause cavitation problems that will interfere with the head.

• The impeller is loose on the shaft. • The impeller is running backwards • The shaft is running backwards because of a wiring problem. • The pump is running backwards because the discharge check valve is not holding

and system pressure is causing the reverse rotation. This is a common problem with pumps installed in a parallel configuration. Check valves are notoriously unreliable.

• The impeller has been installed backwards. This can happen with closed impellers on double ended pumps

• The second stage of a two stage pump is wired backwards. The pump reverses when the second stage kicks in. You should have heard a loud noise when this happened.

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• The wear ring clearance is too large. o This is a common problem if the shaft L3/D4 number is greater than 60 (2

in the metric system). o You should replace the rings when the original clearance doubles.

Needless to say this can only be determined by inspection. • If you are pumping a product at 200°F (100°C) or more you should use a

centerline design volute to prevent excessive wear ring wear as the volute grows from the base straight up, engaging the wear rings.

• A wear ring is missing. It was probably left off during the installation process. • A high suction tank level is reducing the differential pressure across the pump

increasing its capacity. The pump pumps the difference between the suction and discharge heads.

• A bubble is trapped in the eye of the impeller. The eye is the lowest pressure area. When this bubble forms it shuts off all liquid coming into the pump suction. This could cause the pump to lose its prime.

• You cannot vent a running pump because centrifugal force will throw the liquid out the vent leaving the air trapped inside.

• Air is coming directly into the pump. This happens with a negative pressure at the suction side. Negative suction happens when the pump is lifting liquid, pumping from a condenser hot well etc.

o Air is coming into the stuffing box through the pump packing. o Air is coming into the stuffing box through an unbalanced mechanical seal.

As the carbon face wears the spring load holding the faces together diminishes.

o If you are using mechanical seals in vacuum service, they should be of the O-ring design. Unlike other designs, O-rings are the only shape that seals both pressure and vacuum.

o The pump was not primed prior to start up. With the exception of the self priming version, centrifugal pumps must be full of liquid at start up.

o Air can enter the stuffing box if the gasket between the two halves of a double ended pump is defective or does not extend to the stuffing box face. Any small gaps between the face of the stuffing box and the split at the side of the stuffing box will allow either air in, or product out.

o Air is coming into the suction side of the pump through a pin hole in the casing.

o Air is entering the stuffing box between the sleeve and the shaft. This happens if you convert a double ended pump from packing to a mechanical seal and fail to install a gasket or o-ring between the impeller hub and the sleeve.

• The open impeller was adjusted backwards and now the close fitting "pump out vanes" are creating a vacuum in the stuffing box.

• You need a volute casing instead of a concentric casing. Volute casings are much better for producing head.

• You have the wrong size pump. It cannot meet the system curve requirements: • The pump was not selected to meet the system curve requirements because no

system curve was given to the pump supplier.

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• At replacement time the same size pump was purchased because no one had calculated losses in the system.

• The pump was sized from a piping diagram that was thirty five years old. There have been numerous piping changes and additions since the original layout. In many instances additional pumps have been installed and this pump is running in parallel with them, but nobody knows it.

THE PROBLEM IS ON THE SUCTION SIDE OF THE PUMP. THE PUMP COULD BE CAVITATING.

• Air is entering the suction piping at some point. o Air is being pumped into the suction piping to reduce cavitation problems o Fluid returning to the sump is being aerated by too far a free fall. The

return line should terminate below the liquid level. o The fluid is vortexing at the pump inlet because the sump level is too low

and the pump capacity is too high. o Air is coming into the system through valves above the water line or

gaskets in the piping flanges. o The liquid source is being pumped dry. If this is a problem in your

application you might want to consider a self priming pump in the future. • The vapor pressure of the fluid is too close to atmospheric pressure. When it rains

the drop in atmospheric pressure causes the inlet fluid to vaporize. • There is a problem with the piping layout. It is reducing the head on the suction

side of the pump. o There is too much piping between the pump suction and the source tank.

You may need a booster pump or an inducer. The higher the pump speed the bigger the problem.

o There is an elbow too close to the pump suction. There should be at least ten diameters of pipe between the elbow and the pump suction. Suction piping should never run parallel with the pump shaft in a double ended pump installation. This can cause unnecessary shaft thrusting.

o A piece of pipe of reduced diameter has been installed in the suction piping.

o Piping was added on the inlet side of the pump to by-pass a piece of equipment that was installed on the floor.

o A piping to pump reducer has been installed upside down causing an air pocket. Concentric reducers can cause the same problem..

o Multiple pump inlets are too close together. • The pump inlet is too close to the tank floor. • The suction lift is too high. • A gasket with too small an inside diameter has been installed in the suction piping

restricting the liquid flow. • A gasket in the suction piping is not centered and is protruding into the product

stream. • A globe valve has been substituted for a gate valve in the suction piping. The loss

of head in a globe valve is many times that of a gate valve.

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• Two pumps are connected in series. The first pump is not sending enough capacity to the second pump.

• The piping inlet is clogged. • A filter or strainer is clogged or covered with something. • Intermittent plugging of the suction inlet.

o Loose rags can do this. o If the suction is from a pond, river, or the sea, grass can be pulled into the

suction inlet. • A foot valve is stuck. • A check valve is stuck partially closed • The foot valve is too small. • A small clam or marine animal cleared the suction screen, but has now grown

large on the pump side of the screen. • The suction piping diameter has been reduced.

o The suction piping collapsed when a heavy object either hit or ran over the piping.

o Solids have built up on the piping walls. Hard water is a good example of this problem

o A liner has broken away from the piping wall and has collapsed in the piping. Look for corrosion in the piping caused by a hole in the liner.

o A foreign object is stuck in the piping It was left there when the piping was repaired.

o The suction is being throttled to prevent the heating of the process fluid. This is a common operating procedure with fuel pumps where discharge throttling could cause a fire or explosion.

• The pump inlet temperature is too high. o The tank is being heated to deaerate the fluid, but it is heating the fluid up

too much. Look for this problem in boiler feed pump applications. o The sun is heating the inlet piping. The piping should be insulated to

prevent this problem. o The operating temperature of the pumped fluid has been increased to

accommodate the process requirements. o A discharge recirculation line is heating the incoming fluid. You should

direct this line to a reservoir rather than the pump suction. o Steam or some other hot cleaner is being circulated through the lines.

• The problem is in the tank connected to the suction of the pump. o The pump capacity is too high for the tank volume. o The tank float is stuck, showing a higher tank level that does not exist. o The tank vent is partially shut or frozen, lowering the suction pressure. o There is not enough NPSH available for the fluid you are pumping. Maybe

you can use an inducer or booster pump to increase the suction pressure. o A high suction tank level is reducing the differential pressure across the

pump, increasing its capacity and lowering the head.

PROBLEMS ON THE DISCHARGE SIDE OF THE PUMP INCLUDING THE PIPING

• Two pumps are in connected in series. The first pump does not have enough capacity for the second pump. They should be running at the same speed with the same width impeller.

• The pump discharge is connected to the bottom of the tank. The head is low until the level in the tank increases.

• Units in the discharge piping should not normally be shut off, they should be by-passed to prevent too much of a change in the pump's capacity.

• If too many units are being by-passed in the discharge system the head will decrease as the capacity increases. This can happen if an extra storage tank farm is being by-passed because the storage capacity is no longer needed.

• A bypass line has been installed in the pump discharge increasing the capacity and lowering the head.

• Piping or fittings have been removed from the discharge side of the pump reducing piping resistance.

• Connections have been installed in the discharge piping that have increased the demand that increases capacity.

• The pump is acting as an accumulator, coming on when the tank level drops. The head will be low until the accumulator is recharged.

• Consider the possibility of a siphon affect in the discharge piping. This will occur if the pump discharge piping is entering into the top of a tank and discharging at a lower level The pump must build enough head initially to take advantage of the siphoning action.

• A discharge valve (manual or automatic) is opened too much.

The pump is not producing enough capacity to satisfy the application?

Let me begin by pointing out that there are a couple of things you must keep in mind when troubleshooting centrifugal pump problems:





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THE PROBLEM IS IN THE PUMP ITS SELF:

• The impeller diameter is too small • The impeller width is too narrow • The impeller speed is too slow. Check the voltage and frequency • The impeller is damaged. • The impeller is clogged. • The open impeller clearance is too large. • The impeller to cutwater clearance is too large. • The impeller specific speed number is too low. • The impeller has been installed backwards • The shaft is running backwards. • The wear ring clearance is too large. • A wear ring is missing. • The second stage of a two stage pump is wired backwards. • A bubble is trapped in the eye of the impeller. • A low suction tank level is increasing the differential pressure across the pump

decreasing its capacity. • Air is coming into the pump suction through the packing. • Air is coming into the pump suction through an unbalanced mechanical seal. • The pump was not primed prior to star up. • You may need a concentric casing rather than the volute design. • You are using a variable speed motor trying to produce a flat curve. Remember

that both the head and capacity change with speed. • The pump is the wrong size. Someone gave the pump distributor a wrong system

curve

THE PROBLEM IS ON THE SUCTION SIDE OF THE PUMP

• There is too much piping between the pump suction and the source tank. • There is an elbow too close to the pump suction. • A filter or strainer is clogged. • Intermittent plugging of the suction inlet. Loose rags can do this. • A foot valve is stuck • The tank float is stuck. Showing a higher tank level that does not exist. • The tank vent is partially shut or frozen. • A globe valve has been substituted for a gate valve. • A check valve is stuck partially closed • Solids have built up on the piping walls. • A liner has broken away from the piping wall and has collapsed in the piping. • The piping was collapsed by a heavy object that hit the outside of the piping. • A foreign object is stuck in the piping It was left there when the piping was

repaired.

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• A small clam cleared the suction screen, but has now grown large on the pump side of the screen.

• The sun is heating the inlet piping. It should be insulated to prevent this problem. • Piping was added on the inlet side of the pump to compensate for a piece of

equipment that was installed in the shop. • A reducer has been installed upside down. • A discharge recirculation line is heating the incoming fluid. • The pump capacity is too high for the tank volume. • Multiple pump inlets are too close together. • The suction lift is too high. • There is not enough NPSH available for the fluid you are pumping. Maybe you

can use an inducer to increase the suction pressure. • Air is coming into the system through valves above the water line or gaskets in

the piping. • Air is being pumped into the suction piping to reduce cavitation problems • Fluid returning to the sump is being aerated by too far a free fall. • The fluid is vortexing at the pump inlet because the sump level is too low. • The tank is being heated to deaerate the fluid, but it is heating the fluid up too

much. • Two pumps are connected in series. The first pump is not sending enough

capacity to the second pump. • The operating temperature of the pumped fluid has increased. • The vapor pressure of the fluid is too close to atmospheric pressure. When it rains

the drop in atmospheric pressure causes the inlet fluid to vaporize. • The suction is being throttled to prevent the heating of the process fluid.


• Extra piping has been added to the system to accommodate extra storage capacity. • A bypass line has been installed in the pump discharge. • Piping or fittings have been added to the discharge side of the pump. • An orifice has been installed in the discharge piping to reduce the capacity or

produce a false head. • A gate valve has been substituted for a globe valve in the discharge piping. • A check valve is stuck partially closed. • An orifice has been installed into the piping to restrict flow. • The piping was collapsed by a heavy object that hit the outside of the piping. • The discharge valve is throttled too much. • There is a restriction in the discharge piping. • Extra pumps have been installed into the existing piping They are connected in

parallel, but are not producing the same head. • Two pumps are in parallel. The larger one is shutting the check valve of the

smaller pump. • Two pumps are in connected in series. The first pump does not have enough

capacity for the second pump. They should be running at the same speed with the same width impeller

• The pump discharge is connected to the bottom of the tank. The head is increasing and the capacity is decreasing as the tank fills.

• The pump is acting as an accumulator&emdash;coming on when the tank level drops. The head is too high when the tank fills.

Pump selection practices that cause high seal and bearing maintenance problems:

Purchasing well designed hardware does not bring automatic trouble free performance with it. The very best equipment will cause problems if it was not designed for your particular application. Here are a few of the more common selection problems we find with centrifugal pumps:

• Buying the same size pump as the one that came out of the application. That's O.K. If the old pump was the correct size, but the odds are that it was too big because of the safety factors that were added at the time of purchase. This will cause the pump to run off of its best efficiency point (B.E.P.) and you will spend a lot of production money for the additional power that is needed to run against a throttled discharge valve or orifice installed in the discharge piping.

• Buying to a standard, or making a decision based on efficiency, and believing that these two some how relate to quality. Standards were written for packed pumps. When a mechanical seal is being used the shaft L3/D4 number is almost always too large. Efficiency is always gained at the expense of maintenance. Efficiency means tight tolerances and smooth passages that will eliminate reliable double volute designs and keep the maintenance department busy adjusting tight tolerances to maintain the efficiency you paid for.

• Series and parallel installation problems. We often find pumps installed in parallel, but no one knows it because the second pump was installed at a much later date and no one has bothered to trace the piping. Pumps in parallel require that they have the same diameter impeller and that they run at the same speed, or the larger pump will throttle the smaller one causing it to run off the best efficiency point, deflecting the shaft. The capacity should be looked at if the higher capacity pump might exceed the N.P.S.H. available.

• When pumps are installed in series the impellers must be the same width and they must run at the same speed or the higher capacity pump will either cavitate because the smaller capacity pump can not feed liquid at the proper capacity, or it will run throttled if it is feeding the smaller pump. In either case the larger of the two pumps will be adversely affected.

• Purchasing a larger pump because it will be needed in the future. Will raise the operating cost to unacceptable levels (Power = head x capacity) as the pump is run against a throttled discharge valve. This inefficient use of power will translate to a higher heat environment for the seal along with all of the problems associated with shaft deflection.

• Using a variable speed motor to compensate for a pump curve that is not flat enough. Many boiler feed pumps require a flat curve so that the pump can put out

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varying capacities at a constant boiler pressure (head). We see this same need if we are pumping a varying amount of liquid to a very high constant height.

• Varying the speed of a pump is similar to changing the diameter of the impeller. If you look at a typical pump curve you will observe that the best efficiency point (B.E.P.) comes down with impeller size to form an angle with the base line (capacity line) of the graph. This means that if you vary the speed of the impeller, the pump always runs off the B.E.P. except in the case where the system curve intersects the pump curve, or in the case of an exponential system curve such as we find in a typical hot or cold water circulating system.

• Double ended pumps installed in a vertical position to save floor space. Makes seal replacement a nightmare unless you are using split or cartridge designs.

• Specifying a desired capacity without knowing the true system head. You can't guess with this one. Some one has to make the calculations and "walk the system". The present pump is not a reliable guide because we seldom know where it is pumping on its' curve. Chart recorders installed on both the suction and discharge side of the pump will give a more accurate reading of the present head if they are left on long enough to record the differences in flow. The trouble with this method is that it will also record a false head caused by a throttled valve, an orifice, or any other restriction that might be present in the piping.

• Requesting too low a required N.P.S.H. will cause you to end up with a different kind of cavitation problem. See another paper in this series for information about "Internal recirculation".

• Failure to request a "center line design" when pumping temperature exceeds 200°F (100°C) it will cause pipe strain that will translate to wear ring damage and excessive mechanical seal movement.

• The use of "inline" pumps to save floor space. Many of these designs are "close coupled" with the motor bearings carrying the radial and thrust loads. Because of typical L3/D4 numbers being very high, the wear rings act as "steady bearings" after the pump is converted to a mechanical seal. The pump should have been designed with a separate bearing case and a "C" or "D" frame adapter installed to connect a motor to the bearing case.

• Thrust bearings being retained by a simple snap ring. Beyond 65% of its rated efficiency most centrifugal pumps thrust towards the pump volute. The thin snap ring has to absorb all of this axial thrust and most of them can not do it very well .

• The mechanical seal has been installed in a packing stuffing box that is too narrow to allow free seal movement. If a mechanical seal was specified, the pump back plate should have been manufactured with a large diameter seal chamber. In most cases the stuffing box recirculation line should be installed from the bottom of this large seal chamber to the suction side of the pump or a low pressure point in the system. There are some exceptions to this, however:

o If you are pumping at or close to vapor point. o If the entrained solids have a low specific gravity. o If you are using a Duriron pump that adjusts to the back plate. o If you are using a double suction pump where the stuffing boxes are at

suction pressure. • High temperature applications have several special needs:

o A jacketed stuffing box that isolates the pumpage from the stuffing box contents by a carbon bushing to retard heat transfer.

o A centerline design to compensate for thermal expansion. o A cartridge seal design that allows open impeller adjustment after the

pump has come up to operating temperature. o A stainless steel shaft to retard heat transfer to the bearings. o A method of cooling the bearing oil, but never the bearings. o A coupling that will compensate for axial expansion.

The problem with pump standards

All about specific speed How to read a pump curve?

Why prime a centrifugal pump? The oversized pump Understanding the system curve

How efficient is your pump? Piping system for pumps

Calculating the total system head


Prevent potential seal andpump problems

Estimating the shutoff head for pumps

Rules of thumb for pumps Rules of thumb for seals

Calculate the water HP coming out of pump

A new technique of troubleshooting

The best pump and sealtechnology?

Operational practices to avoid problems

Pumps does not develop enough head?

Pump does not give enoughflow?

Pump selection practices leads to problems

OSHA 1910 regulation

The problem with pump standards:

A Quick check of existing pump standards will reveal that there are a variety of them. The list includes:

• Hydraulic Institute Standards • American National Institute Standards for Chemical Pumps :

o B73.1 for Horizontal type. o B73.2 for Vertical Inline

• API 610 for centrifugal Pumps • API 674 for Reciprocating Pumps • API 675 for Controlled Volume Pumps • API 676 for Rotary Positive Displacement Pumps • ISO aimed at the medium duty single stage pumps ( Metric) • DIN. West German standard • VDMA West German standard for pump seals.

There are two problems with these standards:

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• They were written for pumps equipped with jam packing. Most of the standards were written in the nineteen fifties at a time mechanical seals were not popular. In those days we had a lack of the modern materials that make mechanical seals practical. As an example Viton® was not invented until 1958 and did not come into general use until the sixties. Kalrez® did not come out until 1975 and in the eighties the duplex metals came into their own.

• The customer believes that by purchasing a standard design he is getting a good pump. Customers have the same problem with pump efficiency. They believe there is a correlation between efficiency and the quality of the pump, needless to say there is none! Problems caused by these standards are reflected in continual poor seal performance. The fact of the matter is that these standards reflect only an attempt to standardize envelope (outside) dimensions, nothing more!

Unfortunately standardizing the length of pumps prevented manufacturers from designing short shafts that were not prone to the bending problems associated with low cost A.N.S.I. and I.S.O. design pumps, operating off of their best efficiency point (B.E.P).

Here is a list of some of the modifications you should make to your standard A.N.S.I. or I.S.O. pump if you want to get good mechanical seal and bearing life. Unless you are prepared to upgrade the pump seal and bearing life will always be less than desirable

WHAT TO MODIFY

• The stuffing box bore is too small for mechanical seals. In most cases there is not enough material to bore out so you will have to make or purchase a replacement part. Most of these standard stuffing boxes were designed for 3/8" or 10 mm. packing. You need at least 1" (25 mm.) radial clearance to take advantage of centrifugal force throwing solids away from the seal faces.

• When using mechanical seals install a recirculation line from the bottom of the stuffing box back to the suction of the pump. Try to tap the box as close to the face as possible to insure good circulation. Most quality seals come with this connection already installed in the gland.

• Because packing needs lubrication, the pump came equipped with a line from the discharge side to the stuffing box lantern ring connection. If you install a large sealing chamber in place of the narrow packing stuffing box that came as original equipment you should be able to eliminate almost all need for clean flushing liquid in the seal area. The only exception to this is if you are pumping a fluid close to its vaporization point. In that instance you do not want to lower stuffing box pressure because of the possibility of vaporizing the fluid in the stuffing box and possibly blowing open the seal faces

• Convert to Cartridge or Split Seals to insure correct seal installation and allow proper impeller settings in "Back Pull Out" or other types of pump designs.

• If you are using single stage centrifugal pumps convert to solid shafts with a low L3/D4 ratio to resist shaft bending. The back pull out design was made for easy

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sleeve removal. If you are using good mechanical seals, corrosion resistant shaft materials and labyrinth oil seals or positive bearing seals, there should be no need to replace pump shafts.

• Pump manufacturers are not required to provide L3/D4 ratio numbers that would predict shaft bending problems with their pump. The relationship between shaft size and shaft diameter is expressed in the ratio L3/D4. Try to keep it below 60 (2.5 Metric)

o "L" is the distance from the center of the inboard bearing to the center of the impeller (inches).

o "D" is the diameter of the shaft in the stuffing box area.(inches). • Substitute labyrinth or positive face seals for the lip or grease seals that are

installed in the bearing case. They will not only do a better job of keeping contaminants out of the bearing oil/grease, but they will not damage the expensive shaft. These seals also make sense in the motors to eliminate moisture from damaging the windings and contaminating the lubricating grease.

• Use only non- fretting mechanical seals. Shafts are too costly not to pay attention to this.

• The easiest way to get pump/motor alignment is with a "C" (inch) or "D" (metric) frame adapter. If you elect not to use the adapter you are in for a long process aligning the pump and driver correctly, and unless you are using split mechanical seals you are going to have to go through the procedure each time you change seals. You should be able to get the C or D frame adapter as part of your next power frame change or upgrade.

• Convert to a "Center Line" wet end if you are pumping liquids in excess of 200 degrees Fahrenheit (100 Centigrade) It will allow the suction flange to expand without causing pipe strain and wear ring damage.

• Do not use a vent on the bearing cavity of the pump. Each time the pump stops the vent will allow moisture to enter the bearing cavity as the oil cools down (this is called aspiration). You are much better off positively sealing the casing and installing an expansion chamber on the top of the casing to allow for air expansion.

• If you intend to use a closed impeller, end suction, centrifugal pump try to convert to a design that has adjustable wear rings

• Install a sight glass to be sure that the oil level is at the correct height. Too much oil is as bad as not enough. If you have a positive pressure oil mist system be sure that it does not vent to atmosphere. Oil mist systems require mechanical seals outboard of the bearings to prevent atmosphere contamination. If you have installed labyrinth seals, they will almost guarantee the correct oil level because excess oil will spill out of the labyrinth.

• Coat the inside of the bearing case with a suitable protective covering to prevent rusting and the leaching out of harmful substances from the bearing casting.

• Install magnetic plugs into the bottom of the bearing casing to attract loose metal shavings that would damage the bearings.

• Specify double volute designs any time the impeller diameter is 14" (356 mm.) or greater to prevent shaft deflection. Smaller size pumps do not lend themselves to this modification.

• Convert to a "Vortex" pump volute any time you are pumping liquid that contains lots of solids. Although the pump efficiency will be lower than conventional designs, the increased service life will more than compensate.

• At overhaul time substitute a medium or heavy weight power end for the light weight version that came with the pump and get most of the features we have discussed. Medium and heavy weight power ends are available for most popular pump brands.

In addition to these modifications mentioned here are some recommendations to help insure good seal and bearing life.

• Since a seal failure is the most common reason for shutting down a pump, Install a back up seal and convection tank to prevent unexpected shut downs.

• Change the bearing oil on a regular basis. Contact your favorite oil supplier for his recommendation and then follow his advice. If the inside of the bearing frame has been coated with a protective material to prevent rusting, avoid synthetic oils as their detergent action can damage these protective coatings.

• Maintain the proper oil level. Too much is just as bad as not enough. • Trim the impeller to obtain operation at the B.E.P. Throttling the pump discharge

is not the same thing. • If you are using open impellers, keep them adjusted to the correct clearance. • Install bearings by expanding the bore with an induction coil. Heating the bearing

in a pan of warm oil is not a good idea because the oil can easily be contaminated. • Install pressure gages on the suction and discharge of the pump. This is the only

way to tell if the pump is running near its B.E.P. • Do not specify Canned or Magnetic Pumps if the pumping fluid contains solids or

if it is a poor lubricant.

Pumps equipped with a "repeller" and some sort of static seal can usually be converted to a good mechanical seal. The problem with the repeller design is that in most of the designs the seal faces are designed to open when the pump is running and then close on any solids as the pump stops. The rule with mechanical seals is a simple one. "Keep the seal faces together" . Do not open them on purpose.

O.S.H.A. 1910 REGULATION

The regulation is predictably vague, and presently only applies to pressure vessels, storage tanks, processing piping, relief and vent systems, fire protection system components, emergency shut down systems, alarms, interlocks and the part that is important to you, pumps . For the first time Washington is telling the pump user that he has to now document the training he provides to those people (including contractors) that will be operating or repairing his pumps. Be sure to pick up a copy of this regulation for your library.

Here are some of the ingredients you will find in the regulation :

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• The chemicals in the O.S.H.A. # 1910 specification are different than those chemicals identified by the E.P.A. for fugitive emission consideration. The O.S.H.A. list identifies those chemicals that are considered "extremely hazardous" chemicals. O.S.H.A. feels that the general industry standards are not sufficient for these chemicals

• Your employer is going to have to create a Process Safety Management audit team (PSM) that will audit company training programs along with insuring that present and future engineering practices conform to accepted standards and codes.

• The employer is going to have to identify the codes and standards he relied upon to establish his engineering practices. If he departs from these codes and standards, he must document that the design and construction are suitable for the intended purpose.

• The written training programs must be reviewed for adequacy of content, frequency of training, and the effectiveness of the training in terms of goals and objectives. These training programs must be revised if after the training the employee is not at the level of skill or knowledge that was expected.

• Contract employees must also receive updated and current training. • If an accident occurs, the plant is going to have to prove that their training program was

adequate. • Any mechanical changes made by the maintenance department have to be evaluated to

determine whether operating procedures and practices also need to be changed. The term "Change" includes all modifications to equipment.

• For existing processes that have been shut down for turnaround or modification, the employer must ensure that any changes other than "replacement in kind" made to the process during shutdown go through the management of change procedures.

• Equipment installation jobs need to be properly inspected in the field for use of proper materials and procedures to insure that qualified workers do the job.

• The employer must ensure that the contractor has the appropriate job skills, knowledge, and certification.

• The regulations require detailed records of every action taken in maintaining or rebuilding a pump. The employer must identify which procedures were followed and why he elected to use those procedures. He must also identify the training that maintenance personnel had on repairing pumps in that service.

• Equipment used to process, store or handle hazardous chemicals has to be designed, constructed, installed and maintained to minimize the risk of release of such chemicals.

• The employer must prepare three lines of defense to prevent hazardous chemical from injuring personnel:

o Contain the chemical in the equipment. The use of two mechanical seals and a convection tank is a good example of containing the chemical.

o Control the release of the chemicals through venting with a seal quench and vent connection to a scrubber or flare, or to surge or overflow tanks designed to receive such chemicals. Dikes or designed drainage systems would be another alternative.

o A sensible evacuation system is the third line of defense.

If an accident happens and any of the listed chemicals are released to the environment, the employer is going to have to prove he did every thing he could have to prevent the accident and contain the spill. If O.S.H.A. does not agree with his assessment, the employer is likely to suffer stiff penalties.

Since you have knowledge that 90% of mechanical seals are failing prematurely (the carbon sacrificial face is not wearing out) I expect this new regulation should encourage your employer to send more people to seal and pump schools and enroll his engineering, maintenance, and supervisory people in an appropriate certification training program.

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Why you must prime a centrifugal pump

Although the term "pressure" is not normally a part of a centrifugal pump man's vocabulary, we are going to have to discuss it for a couple of minutes.

The earth's atmosphere extends approximately fifty miles (80 Km.) above the earth, and rests on the earth with a weight equivalent to a layer of fresh water thirty four feet (10 meters) deep at sea level. To remove air from the pump cavities and the suction piping, the pump must develop enough head to equal the equivalent of this 14.7 psi., or one bar pressure. In an earlier paper we learned how to convert this height (head) to a pressure reading by use of the following formulas:

34 Ft./2.31 = 14.7 Psi.

10 Meters/10 = 1 Bar

Unlike a positive displacement pump that can pump a liquid to any head as long as the pump body is strong enough, and there is enough horsepower available, the centrifugal pump can only pump a liquid to its rated head. You will recall that this head was determined by, and limited to the diameter of the impeller and the impeller speed (rpm.)

Since the weight of water is approximately 8000 times that of air (50 miles vs. 34 feet or 80 Km. vs. 10 meters) the centrifugal pump can produce only 1/8000 of its rated liquid pressure. In other words, for every one foot water has to be raised to prime the pump, the centrifugal pump must produce a discharge head of approximately 8000 feet (each meter requires a head of 8000 meters) and that is impossible with conventional impeller diameters and speeds.

All of this means that if you intend to use a centrifugal pump you are going to have to come up with some sensible method of priming it. Your choices will include :

• Install a foot valve in the suction piping to insure the liquid will not drain from the pump casing and suction piping. Keep in mind that these valves have a nasty habit of leaking.

• Evacuate the air in the system with a positive displacement priming pump operating between the pump and a closed discharge valve.

• Fill the pump with liquid prior to starting it. • Convert the application to a self priming pump that maintains a reservoir of liquid

at its suction.

How efficient is your pump?

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A few years ago, efficiency became "the name of the game". Automotive companies advertised "miles per gallon (liters per 100 kilometers) information in their advertisements and appliance manufacturers published kilowatt consumption numbers along with their pricing information.

Unfortunately high efficiency also means higher maintenance costs because you are required to maintain tighter tolerances and keep the flow passages smooth and free from obstructions. The demise of the double volute pump design in smaller size pumps, is a perfect example of the increase in mechanical seal problems as the efficiency of the volute pump was increased to satisfy consumer demand.

Maybe the "trade off" is acceptable as long as you are dealing with accurate numbers, but are you really doing that? Is the efficiency shown on the pump curve accurate? How was the data taken? What was included in the data, and more important, what was left out? As an example:

• Was the data generated on a dynamometer with a constant speed motor? • Are you going to run at the same speed as shown on the performance curve, or are

you running with an induction motor that slips 2% to 5% and you are not sure of the actual speed? Horse power (K.W.) varies as the cube of the change in speed at the best efficiency point, so a small variation in speed can make a big difference in efficiency.

• Was the published efficiency data generated with a seal or packing in the stuffing box? The type of packing or seal used can alter the load they consume.

• Was there an elbow at the suction of the pump? • Was the inside of the volute polished or coated with a low friction material when

the test was made? • How were the bearings lubricated, and were all of the losses considered in the

published numbers? • The final numbers will vary with the motor efficiency, and that will vary with the

load on the motor.

If you would like to keep the pump salesman honest, take the data from his pump curve and then make the following calculation:

In inch sizes : GPM x TDH / 3960 = WHP

• GPM = Gallon per minute at the best efficiency point • TDH = Total discharge head (measured in feet), as shown on the pump

curve&emdash; at the best efficiency point) • WHP = Water horse power, or the amount of horse power the pump is generating.

If we refer to the above pump curve, and insert the numbers into our formula, we would get:

in inch size: GPM x TDH / 3960 = 250 X 300 / 3960 = 18.9

You then divide this number by the efficiency shown on the pump curve:

18.9 / .60 = 31.5 horsepower required to generate the WHP. If this number is lower than the horsepower shown on the performance curves, the efficiency date is questionable. As an example:

If the performance curve showed a requirement for 40 Horse power, the actual efficiency would be 18.9 water horse power40 pump horsepower = .47 or 47% actual efficiency.

Doing the same thing in the metric system we would get:

• M3/ HR = Cubic meters per hour of capacity as measured at the best efficiency point on the pump curve.

• TDH = Total discharge head, in meters, at the best efficiency point. • WKW = Water kilowatts of power being generated by the pump.

Referring to the above diagram, and putting in the numbers :

M3 / HR X TDH / 360 = 68 x 76 / 360 = 14.36 WKW. The curve shows a 60% efficiency so:

14.36 water kilowatts / 0.60 efficiency = 23.93 Kilowatts required. If this number is lower than shown on the pump performance curve, the efficiency of the plump is questionable. As an example:

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If the pump performance curve showed a requirement for a 30 Kilowatt input, the actual efficiency would be:

14.36 water horse power / 30 Kilowatts required = 48 % actual efficiency.

The fact of the matter is that you seldom operate at the best efficiency point so the numbers become even more depressing. The point is that efficiency should only be one of the points taken into consideration when you purchase a centrifugal pump of a given head, material and capacity. Equally if not more important should be:

• The L3/D4 number of the shaft. Is the number below 60 in inch sizes or 2 in metric? • What kind of mechanical seal is installed? Will it seal fugitive emissions? • How are the bearings being lubricated? • How are the bearings sealed? Will the seal damage the expensive shaft? • How is the thrust bearing being retained? In operation the impeller thrusts towards

the volute. Are you relying upon a simple snap ring? • Is the pump a centerline design? It should be if the product you are pumping is

greater than 200°F (100 C.) • Is the bearing case vented to atmosphere? If it is, it will allow moisture to

penetrate when the pump stops. • Has a "C" or "D" frame adapter been installed to reduce alignment time? • Can the wear rings or open impeller be easily adjusted to compensate for normal

wear so that you can keep the efficiency you paid for? • Can the seal compensate for thermal growth, or impeller adjustment?

You can save money by lowering operating costs (efficiency) or increasing the time between repairs (design). Be sure you consider both when you make your pump buying decision.


When we talk about automobiles and discuss efficiency, we mean how many miles per gallon, or liters per 100 kilometers. When we discuss centrifugal pumps we are comparing the amount of work or power we get out of the pump to the amount of power we are putting into the pump. As an example:

How do we measure the horsepower or kilowatts coming out of the pump? All we have to do is multiply the pump head by the weight of the liquid being pumped, and then use a simple conversion number. Let's take an example:

Flow = 300 gallons per minute of fresh water as measured coming from the pump discharge.

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Head = 160 feet. We measured it at the discharge side of the pump and corrected it for the fact that the gage was two feet above the pump center line. Look at the following diagram where we have calculated the discharge head from the formula shown on the right hand side of the illustration. If there were any positive head on the suction side of the pump that head would have to be subtracted. A negative suction head would be added to the discharge head.

The centrifugal pump pumps the difference between the suction and the discharge heads. There are three kinds of discharge head:

• Static head. The height we are pumping to, or the height to the discharge piping outlet that is filling the tank from the top. Note: that if you are filling the tank from the bottom, the static head will be constantly changing.

• Pressure head. If we are pumping to a pressurized vessel (like a boiler) we must convert the pressure units (psi. or Kg.) to head units (feet or meters).

• System or dynamic head. Caused by friction in the pipes, fittings, and system components. We get this number by making the calculations from published charts ( non included in this paper, but available in the chart section of this web site).

Suction head is measured the same way.

• If the liquid level is above the pump center line, that level is a positive suction head. If the pump is lifting a liquid level from below its center line, it is a negative suction head.

• If the pump is pumping liquid from a pressurized vessel, you must convert this pressure to a positive suction head. A vacuum in the tank would be converted to a negative suction head.

• Friction in the pipes, fittings, and associated hardware is a negative suction head. • Negative suction heads are added to the pump discharge head, positive suctions

heads are subtracted from the pump discharge head.

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Here is the formula for measuring the horsepower out of the pump:

Remember that we are using the actual horsepower or kilowatts going into the pump and not the horsepower or kilowatts required by the electric motor. Most motors run some where near 85% efficient.

An 85% efficient motor turning a 76% efficient pump, gives you a real efficiency of 0 .85 x 0.76 = 0 .65 or 65% efficient.

A survey of popular pump brands demonstrates that pump efficiencies range from 15% to over 90%. The question then arises, "Is this very wide range due to poor selection, poor design, or some other variable which would interfere with good performance?" The best available evidence suggests that pump efficiency is directly related to " the specific speed number " with efficiencies dropping dramatically below a number of 1000 . Testing also shows that smaller capacity pumps exhibit lower efficiencies than higher capacity designs.

Now that we have learned that pump efficiency is closely related to the shape of the impeller, and the impeller shape is usually dictated by the operating

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conditions, you should be aware of various conditions that decrease the efficiency of your pump. These would include:

• Packing generates approximately six times as much heat as a balanced mechanical seal.

• Wear rings and impeller clearances are critical. Anything that causes these tolerances to open will cause internal recirculation that is wasting power as the fluid is returned to the suction of the pump. If the wear ring is rubbing, the generated heat is consuming power.

• A bypass line installed from the discharge side of the pump to the suction piping. The heat generated from this recirculation can, in some cases, cause pump cavitation as it heats the incoming liquid.

• A double volute design pump restricts the discharge passage lowering the overall efficiency.

• Running the pump with a throttled discharge valve. • Eroded or corroded internal pump passages will cause fluid turbulence. • Any restrictions in the pump or piping passages such as product build up, a

foreign object, or a stuck check valve. • Over lubricated or over loaded bearings. • Rubbing is a major cause. It can be caused by:

o Misalignment between the pump and driver. o Pipe strain. o Impeller imbalance. o A bent shaft. o A close fitting bushing. o Loose hardware. o A protruding gasket rubbing against the mechanical seal. o Cavitation. (5 kinds) o Harmonic vibration. o Improper assembly of the bearings, seal, wear rings, packing, lip seals etc.. o Thermal expansion of various components in high temperature

applications. The impeller can hit the volute, the wear rings can come into physical contact etc.

o Solids rubbing against the rotating components, especially the seal. o Operating too far off of the best efficiency point of the pump. o Water hammer and pressure surges. o Operating at a critical speed. o Dynamic, non O-ring elastomers that cannot flex and roll, but must slide,

eventually fretting the shaft or sleeve. o A build up of product on the inside of the stuffing box rubbing against the

mechanical seal. o Grease or lip seals rubbing the shaft next to the bearings. o Over tightening packing or improper seal installation.

• Vortex pumps can lower efficiency by as much as 50%.

All about specific speed

Specific speed is a term used to describe the geometry (shape) of a pump impeller. People responsible for the selection of the proper pump, for their application, can use this Specific Speed information to :

• Select the shape of the pump curve. • Determine the efficiency of the pump. • Anticipate motor overloading problems. • Predict N.P.S.H. requirements. • Select the lowest cost pump for their application.

Specific speed is defined as "the speed of an ideal pump geometrically similar to the actual pump, which when running at this speed will raise a unit of volume, in a unit of time through a unit of head".

The performance of a centrifugal pump is expressed in terms of pump speed, total head, and required flow. This information is available from the pump manufacturer's published curves. Specific speed is calculated from the following formula, using data from these curves at the pump's best efficiency point (B.E.P.):

N = The speed of the pump in revolutions per minute (rpm.)

Q = The flow rate in liters per minute ( for either single or double suction impellers)

H = The total dynamic head in meters

Please refer to the following chart:

Pumps are traditionally divided into three types: radial flow, mixed flow, and axial flow. When you look at the above chart you can see there is a gradual change

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from the radial flow impeller, which develops pressure principally by the action of centrifugal force, to the axial flow impeller, which develops most of its head by the propelling or lifting action of the vanes on the liquid.

In the specific speed range of approximately 1000 to 6000 double suction impeller are used as frequently as the single suction impellers.

If you substitute other units for flow and head the numerical value of Ns will vary. The speed is always given in revolutions per minute (rpm.). Here is how to alter the Specific Speed number (Ns) if you use other units for capacity and head :

• United States ....Q = G.P.M. and H = feet. Divide the Ns by 1.63 • British ............Q = Imp.G.P.M. and H = feet. Divide the Ns by 1.9 • Metric ............Q = M3/hour and H = meters. Divide the Ns by 1.5

As an example we will make a calculation of Ns in both metric and U.S. units :

• Q= 110 L/sec. or 396 M3/ hour or 1744 G.P.M. • H = 95 meters or 312 feet • Speed = 1450 rpm.

If the above results were describing an actual application, we would notice that it was a low specific speed, radial flow pump, meaning It would be a large pump with a low efficiency. Going to 2900 rpm. or higher would increase the Ns to 1000 or more, meaning a smaller pump with a much higher efficiency, but this higher rpm. would have other possible consequences :

• The higher efficiency would allow you to use a less powerful driver that would reduce your operating costs.

• A smaller pump makes associated hardware cheaper. For instance, a smaller diameter shaft means a lower cost mechanical seal and lower cost bearings.

• Cavitation could become a problem as the increase in speed means an increase in the N.P.S.H. required.

• If you are pumping an abrasive fluid, abrasive wear and erosion will increase with increasing speed.

• Many single mechanical seals have problems passing fugitive emission standards at the higher pump speeds.

• High heat is a major cause of bearing failure. The higher pump speeds contribute to the problem.

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The following diagram illustrates the relationship between specific speed and pump efficiency. In general, the efficiency increases as Ns increases.

Specific speed also relates to the shape of the individual pump curve as it describes head, capacity, power consumption and efficiency.

In the above diagram you will note that :

• The steepness of the head/ capacity curve increases as specific speed increases. • At low specific speed, power consumption is lowest at shut off and rises as flow

increases. This means that the motor could be over loaded at the higher flow rates unless this was considered at the time of purchase.

• At medium specific speed the power curve peaks at approximately the best efficiency point. This is a non overloading feature meaning that the pump can work safely over most of the fluid range with a motor speed to meet the B.E.P. requirement.

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• High specific speed pumps have a falling power curve with maximum power occurring at minimum flow. These pumps should never be started with the discharge valve shut. If throttling is required a motor of greater power will be necessary.

Keep in mind that efficiency and power consumption were calculated at the best efficiency point (B.E.P.). In practice most pumps operate in a throttled condition because the pump was oversized at the time it was purchased. Lower specific speed pumps may have lower efficiency at the B.E.P., but at the same time will have lower power consumption at reduced flow than many of the higher specific speed designs.

The result is that it might prove to be more economical to select a lower specific speed design if the pump had to operate over a broad range of capacity.

The oversized pump

Do a survey of any process plant and you will find that a high percentage of the centrifugal pumps are oversized. There must be a reason why this is such a common problem, so here are a few of them :

• Safety margins were added to the original calculations. Several people are involved in the pump buying decision and each of them is afraid of recommending a pump that proves to be to small for the job.

• It was anticipated that a larger pump would be needed in the future, so it was purchased now to save buying the larger pump later on.

• It was the only pump the dealer had in stock and you needed one badly. He might have offered you a "special deal" to take the larger size.

• You took the pump out of your spare parts inventory. Capital equipment money is scarce so the larger pump appeared to be your only choice.

• You purchased the same size pump as the one that came out of the application and that one was over sized also.

Obviously this larger pump and motor required a higher investment, but since we are not using the full power are we really paying too much for the daily operation? The easiest way to find the answer to this question is to look at a typical pump curve and make our calculations from the numbers we get.

You can use any of the following formulas to make your calculations:

Here is as typical pump curve. It can be used for both inch and metric examples.

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Let us assume that the application requires a pump that moves the liquid at :

300 gpm. to a 156 foot head with an efficiency rating of 60%

156 x 300 / 5308 = 8.8 Kilowatts being produced, and 8.8 / 0.60 efficiency = 14.7 Kilowatts required

As shown in the above drawing, we should be using impeller "E" to do this, but we have an oversized pump so we are using the larger impeller "A" with the pump discharge valve throttled back to 300 gpm. giving us an actual head of 250 feet and a 50% efficiency. Now our Kilowatts look like this:

250 x 300 / 5308 = 14.1 KW being produced, and 14.1 / 0.50 efficiency = 28.2 KW

required to do this. If 28.2 KW is being used and only 14.7 KW are required, it means that we are paying for an extra 13.5 KW to pump against the throttled discharge valve.

If this pump runs 24 hours per day that would be 8760 hours this year, and at a power cost of $0.05 cents per Kilowatt hour it would cost your company an additional:

8760 hours. x .05 cents per Kilowatt hour x 13.5 Kilowatts = $5913.00 per year, extra operating cost.

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Now we will work the same problem in the metric system:

Assume that we need to pump 68 m3/hr. to a 47 meter head with a pump that is 60% efficient at that point.

68 x 47360 = 8.9 Kilowatts being produce, and 8.9 / 0.60 efficient = 14.8 Kilowatts required to do this.

As shown in the drawing, we should be using impeller "E" to do this, but we have an oversized pump so we are using the larger impeller "A" with the pump discharge valve throttled back to 68 cubic meters per hour, giving us an actual head of 76 meters. Now our Kilowatts look like this:

68 x 76360 = 14.3 Kilowatts being produced by the pump, and 14.3 / 0.50 efficient = 28.6 Kilowatts required to do this.

Subtracting the amount of kilowatts we should have been using gives us:

28.6 - 14.8 = 13.8 extra kilowatts being used to pump against the throttled discharge valve. If the pump runs twenty four hours a day that would be 8760 hours per year, times 13.8 extra kilowatts equals 120,880 kw. Multiply this number by how much you spend for a kilowatt hour of electricity and you will see that the over sized pump is costing you a lot of money. In this example the extra cost of the electricity could almost equal the cost of purchasing the pump.

How to read a pump curve

Please look at the above illustration. You will note that I have plotted the head of the pump against its capacity. The head of a pump is read in feet or meters. The capacity units will be either gallons per minute, liters per minute, or cubic meters per hour.

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According to the above illustration this pump will pump a 40 capacity to about a 110 head, or a 70 capacity to approximately a 85 head (you can substitute either metric or imperial units as you see fit)

The maximum head of this pump is 115 units. This is called the maximum shutoff head of the pump. Also note that the best efficiency point (BEP) of this impeller is between 80% and 85% of the shutoff head. This 80% to 85% is typical of centrifugal pumps, but if you want to know the exact best efficiency point you must refer to the manufacturers pump curve.

Ideally a pump would run at its best efficiency point all of the time, but we seldom hit ideal conditions. As you move away from the BEP the shaft will deflect and the pump will experience some vibration. You will have to check with your pump manufacturer to see how far you can safely deviate from the BEP (a maximum of 10% either side is typical)

Now look at the following illustration:

Note that I have added some additional curves to the original illustration. These curves show what happens when you change the diameter of the impeller.

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Impeller diameter is measured in either inches or millimeters. If we wanted to pump at the best efficiency point with a 11.5 impeller we would have to pump a capacity of 50 to a 75 head.

The bottom half of the illustration shows the power consumption at various capacities and impeller diameters. I have labeled the power consumption horsepower, but in the metric system it would be called kilowatts

Each of the lines represents an impeller diameter. The top line would be for the 13 impeller the second for the 12.5 etc. If we were pumping a capacity of 70 with a 13 impeller it would take about 35 horsepower. A capacity of 60 with the 12 impeller would take about 20 horsepower.

Most pump curves would show you the percent of efficiency at the best efficiency point . The number varies with impeller design and numbers from 60% to 80% are normal.

When you will look at an actual pump curve you should have no trouble reading the various heads and corresponding capacities for the different size impellers. You will note however, that the curve will usually show an additional piece of information and that is NPSHR which stands for net positive suction head required to prevent the pump from cavitating.

Depending upon the pump curve you might find a 10 foot (3.0 meter) NPSH required head at a capacity of 480 Gallons per minute (110 cubic meters per hour) if you were using a 13 inch (330 mm.) diameter impeller.

You should keep in mind that the manufacture assumed you were pumping 20° C ( 68° F ) fresh water and the N.P.S.H. Required was tested using this assumption. If you are pumping water at a different temperature or if you are pumping a different fluid, you are going to have to add the vapor pressure of that product to the N.P.S.H. Required. The rule is that Net Positive Suction Head Available minus the Vapor Pressure of the product you are pumping (converted to head) must be equal to or greater than Net Positive Suction Head Required by the manufacturer.

Suppose we wanted to pump some liquid Butane at 32 degrees Fahrenheit (0 degrees Centigrade) with this pump. If we look at the curve for Butane on a vapor pressure chart similar to the one shown in the charts and graphs section of this web site you will note that Butane at 32°F needs at least 15 psi (1,0 Bar) to stay in a liquid state. To convert this pressure to head we use the standard formula :

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In other words Butane at this temperature would not vaporize as long as I had the above absolute heads available at the suction side of the pump.

Understanding the system curve

Every pump manufacturer would like to recommend the perfect pump for your application. To do this he would like you to provide him with an accurate system curve that would describe the capacity and head needed for your various operating conditions. Once he has your system curve, he can plot his pump curves on top of the system curve and hopefully select something that will come close to your needs. Without this system curve, neither one of you has much of a chance of coming up with the right pump.

To create a system curve we plot the desired capacities against the required head over the total anticipated operating range of the pump. The head will be measured in feet or meters and the capacity will be measured in gallons per minute or cubic meters per hour.

Some of the confusion begins when we realize that there are three different kinds of head:

STATIC HEAD This is the vertical distance measured from the center line of the pump to the height of the piping discharge inside the tank. Look at figure "A" and note that the piping discharge is below the maximum elevation of the piping system. We do not use the maximum elevation in our calculations because the siphoning action will carry the fluid over this point once the piping is full of liquid. This is the same action that lets you siphon gasoline out of an automobile to a storage can.

The pump will have to develop enough head to fill the pipe and then the siphoning action will take over. The pump operating point should move back towards the best efficiency point (B.E.P.) if the pump was selected correctly.

FIGURE "A"

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DYNAMIC OR SYSTEM HEAD As the liquid flows through the piping and fittings, it is subject to the friction caused by the piping inside finish, restricted passages in the fittings and hardware that has been installed in the system. The resulting "pressure drop" is described as a "loss of head" in the system, and can be calculated from graphs and charts provided by the pump and piping manufacturers. These charts are not included with this paper, you can find them in the Hydraulic Institute Manuals. This "head" loss is related to the condition of the system and makes the calculations difficult when you realize that older systems may have "product build up" on the piping walls, filters, strainers, valves, elbows, heat exchangers, etc., making the published numbers some what inaccurate.

A general "rule of thumb" says that the friction loss in clean piping will vary approximately with 90% of the square of the change in flow in the piping, and 100% of the square with the change of flow in the fittings and accessories. You calculate the change in flow by dividing the new flow by the old flow and then square the number. As an example:

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In the original application system, loss was a combination of the loss through the piping and the loss through the fittings for a total of 100 feet at 200 gallons per minute. When we increased the flow to 300 gallons per minute our system head changed to a total of 208.13 feet. This change would have to be added to the static and pressure heads to calculate the total head required for the new pump.

Please note that the pump is pumping the difference between the suction head and the discharge head, so if you fail to consider that the suction head will be either added to or subtracted from the discharge head, you will make an error in your calculations. The suction head will be negative if you are lifting liquid from below ground or if you are pumping from a vacuum. It will be positive if you are pumping from a tank located above ground. If the suction head is pressurized, this pressure must be converted to head and subtracted from the total head required by the pump.

A centrifugal pump will create a head/capacity curve that will generally resemble one of the curves described in figure "B" The shape of the curve is determined by the Specific Speed number of the impeller.

Centrifugal pumps always pump somewhere on their curve, but should be selected to pump as close to the best efficiency point (B.E.P.) as possible. The B.E.P. will fall some where between 80% and 85% of the shut off head (maximum head).

The manufacturer generated these curves at a specific R.P.M.. Unless you are using synchronous motors (you probably are using induction motors on your pumps) you will have to adjust the curves to match your actual pump speed. Put a tachometer on the running motor and record the rpm. difference between your pump and the speed shown on the pump manufacturer's published curve. You can use the pump affinity laws to approximate the change.

POSITIVE DISPLACEMENT PUMPS have a different shaped curve. They look something like Figure "C".

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In this system, the head remains a constant as the capacity varies. This is a typical application for:

• A boiler feed pump that is supplying a constant pressure boiler with a varying steam demand. This is a very common application in many process systems or aboard a ship that is frequently changing speeds (answering bells).

• Filling a tank from the top and varying the amount of liquid being pumped, is the normal routine in most process plants. The curve will look like this if the majority of the head is either static or pressure head.

The second system is the ideal one, Figure "E" describes it:

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In this system the entire head is system head so it will vary with the capacity. Look for this type of curve in the following applications:

• A circulating hot or cold water heating/ cooling system. • Pumping to a non pressurized tank, a long distance from the source with little to

no elevation involved. Filling tank cars is a typical application.

System curve "G" is a common one. It is a combination of static, pressure and system heads.

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Once the pump manufacturer has a clear idea as to the shape of your system curve, and the head and capacity numbers needed he can then select the proper centrifugal pump. The shape of his curve will be pretty much determined by the specific speed number of the impeller.

In addition to specific speed he can select impeller diameter, impeller width, pump rpm., and he also has the option of series or parallel operation along with the possibility of using a multi-stage pump to satisfy your needs.

The sad fact is that most pumps are selected poorly because of the desire to offer the customer the lowest possible price. A robust pump, with a low L3/D4, is still your best protection against seal and bearing premature failure when the pump is operating off of its best efficiency point. Keep the following in mind as you select your pump:

• A centrifugal pump will pump where the pump curve intersects the system curve. This may bear no relationship to the best efficiency point (B.E.P.), or your desire for the pump to perform a specific task.

• The further off the B.E.P. you go, the more robust the pump you will need. This is especially true if you have replaced the packing with a mechanical seal and no longer have the packing to act as a support bearing when the shaft deflects. Shaft deflection is always a major problem at start up.

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• When you connect pumps in parallel, you add the capacities together. The capacity of a pump is determined by the impeller width and r.p.m.. The head of a centrifugal pump is determined by the impeller diameter and rpm. If the heads are different, the stronger pump will throttle the weaker one, so the impeller diameters and rpm's must be the same if you connect pumps in parallel. Check the rpm's on these pumps if you are experiencing any difficulties.

• If you connect the pumps in series, the heads will add together, so the capacities must be the same or one of them will probably cavitate. You could also have a problem operating too far to the right of the best efficiency point with a possible motor "burn out".

• When you vary the speed of a centrifugal pump, the best efficiency point comes down at an angle. The affect is almost the same as changing the diameter of the impeller. This means that the variable speed motor will work best on a system curve that is exponential (Figure "F"). Unfortunately most process and boiler feed pump system curves are not exponential.

• Pump curves are based on a speed of 1750, 3500, 1450, or 2900 r.p.m.. Electric induction motors seldom run at these speeds because of "slip". You can estimate that a 2% to a 5% slip is normal in these pumps with the "slip" directly related to the price of the motor.

• You should also keep in mind that if the motor is running at its best efficiency point that does not mean that the pump is running at its B.E.P..

Since you will be using pumps that were supplied at the lowest cost, you can do the following to resist some of the shaft displacement:

• Use a solid shaft. Sleeves often raise the L3/D4 number to over 60 (2 in the metric system), and this is too high a number for reliable seal performance.

• Try to keep the mechanical seal as close to the bearings as possible. It is the mechanical seal that is the most sensitive to shaft deflection and vibration.

• Once the seal has been moved closer to the bearings, you can install a sleeve bearing in the packing space to support the shaft when the pump is operated off of its B.E.P. This is especially important at start up, or any time a pump discharge valve is operated.

• Stop the cavitation if you are experiencing any. • Balance the rotating assembly. • Check that the shaft is not bent or the rotating assembly is not out of dynamic

balance. • Use a "C" or "D" frame adapter to solve pump- motor alignment difficulties. • A center line design wet end can be used if pipe strain, due to temperature

expansion, is causing an alignment problem.

Do not trust the system prints to make your calculations. The actual system always differs from that shown on the print, because people tap into the lines, using the pumped fluid for a variety of purposes and after having done so forget to change or "mark up" the original system print. You are going to have to "walk down" the system and note the pipe length, the number of fittings, etc., to make

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an accurate system head calculation. Do not be surprised to find that the discharge of your pump is hooked up to the discharge of another pump further down the line. In other words, the pumps are connected in parallel and no body knows it. Pressure recorders (not gauges) installed at the pump suction and discharge is another technique you can use to get a better picture of the system or dynamic head. They will show you how the head is varying with changes in flow.

Pump selection is simple but not easy. Do not depend upon the knowledge of the local pump salesman to select the correct pump for you. In many cases he is prepared to sell his pump at cost&emdash;to get the spare parts business. If you are purchasing pumps at too big a discount&emdash;something is wrong, there is no free lunch. Keep in mind that if several people are involved in the selection process each of them will commonly add a safety factor to the calculated pump size. These factors added together can cause you to purchase a pump that is very much over sized.

Calculating the total system head in USCS units

USCS stands for "United States Customary System Units" as opposed to the SI (Le Syst`eme International d`Units) or metric units that have been adopted by the International standards Organization (ISO). In a future paper I will present another paper using the metric units, but for the moment it is not convenient to present it in both systems.

It turn out that "head" is a very convenient term in the pumping business. Capacity is measured in gallons per minute, and each gallon of liquid has weight, so we can easily calculate the pounds per minute being pumped. Head or height is measure in feet, so if we multiply these two together we get foot- pounds per minute which converts directly to work at the rate of 33,000 foot pounds per minute equals one horsepower.

Pressure is not as convenient a term because the amount of pressure that the pump will deliver depends upon the weight (specific gravity) of the liquid being pumped and the specific gravity changes with temperature, fluid, and fluid concentration.

If you will refer to FIG 1, you should get a clear picture of what is meant by static head. Note that we always measure from the center line of the pump to the highest liquid level

To calculate head accurately we must calculate the total head on both the suction and discharge sides of the pump. In addition to the static head we will learn that there is a head caused by resistance in the piping, fittings and valves called friction head, and a head caused by any pressure that might be acting on the liquid in the tanks including atmospheric pressure, called " surface pressure head".

Once we know these heads it gets simple, we will then subtract the suction head from the discharge head and the amount remaining will be the amount of head that the pump must be able to generate at the rated flow. Here is how it looks in a formula:

System head = total discharge head - total suction head

H = hd - hs

The total discharge head is made from three separate heads:


• hd = total discharge head • hsd = discharge static head • hpd = discharge surface pressure head • hfd = discharge friction head

The total suction head also consists of three separate heads

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hs = hss + hps - hfs

• hs = total suction head • hss = suction static head • hps = suction surface pressure head • hfs = suction friction head

As we make these calculations, you must sure that all calculations are made in either "feet of liquid gauge" or "feet of liquid absolute". In case you have forgotten "absolute means that you have added atmospheric pressure (head) to the gauge reading.

Now we will make some actual calculations:

Figure #2 demonstrates that the discharge head is still measured to the liquid level, but you will note that it is below the maximum height of the piping.

Although the pump must deliver enough head to get up to this maximum piping height, it will not have to continue to deliver this head when the pump is running because of the "siphon effect". There is of course a maximum siphon effect. It is derived from: 14.7 psi (atmospheric pressure) x 2.31 feet / psi = 33.4 feet maximum siphon effect.

We will begin with the total suction head calculation

1. The suction head is negative because the liquid level in the suction tank is below the centerline of the pump:

hss = - 6 feet

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2. The suction tank is open, so the suction surface pressure equals atmospheric pressure :

hps = 0 feet gauge

3. You will not have to calculate the suction friction head, I will tell you it is:

hfs = 4 feet at rated flow

4. The total suction head is a gauge value because atmosphere was given as 0,

hs = hss + hps - hfs = -6 +0 -4 = -10 feet of liquid gauge at rated flow

The total discharge head calculation

1. The static discharge head is:

hsd = 125 feet

2. The discharge tank is also open to atmospheric pressure, thus:

hpd = 0 feet, gauge

3. I will give you the discharge friction head as:

hfd = 25 feet at rated flow

4. The total discharge head is:

hd = hsd + hpd + hfd = 125 + 0 + 25 = 150 feet of liquid gauge at rated flow

The total system head calculation:

H = hd - hs = 150 - (-10)= 160 feet of liquid at rated flow

Note: did you notice that when we subtracted a minus number (-10) from a positive number (150) we ended up with a positive 160 because whenever you subtract minus numbers it is the same as adding them? If you have trouble with this concept you can learn more about it from a mathematics book.

Our next example involves a few more calculations, but you should be able to handle them. In this example we are going to learn how to handle a vacuum application. Pipe friction numbers are taken from the Hydraulic Institute Engineering Data Book. You can get a copy of this publication from your library if you want to see the actual charts. I have some of this information in the chart section of this web site.

Specifications:

1. Transferring 1000 gpm. weak acid from the vacuum receiver to the storage tank

2. Specific Gravity - 0.98

3. Viscosity -equal to water

4. Piping - All 6" Schedule 40 steel pipe

5. Discharge piping rises 40 feet vertically above the pump centerline and then runs 400 feet horizontally. There is one 90° flanged elbow in this line

6. Suction piping has a square edge inlet, four feet of pipe, one gate valve, and one 90° flanged elbow all of which are 6" in diameter.

7. The minimum level in the vacuum receiver is 5 feet above the pump centerline.

8. The pressure on top of the liquid in the vacuum receiver is 20 inches of mercury, vacuum.

To calculate suction surface pressure use one of the following formulas:

• inches of mercury X 1.133specific gravity = feet of liquid • pounds per square inch X 2.31specific gravity = feet of liquid

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• Millimeters of mercury X 122.4 x specific gravity = feet of liquid

Now that you have all of the necessary information we will begin by dividing the system into two different sections, using the pump as the dividing line.

Total suction head calculation

1. The suction side of the system shows a minimum static head of 5 feet above suction centerline. Therefore, the static suction head is:

hss = 5 feet

2. Using the first conversion formula, the suction surface pressure is:

hps = -20 Hg X 1.133/ 0.98 = -23.12 feet gauge

3. The suction friction head, fs, equals the sum of all the friction losses in the suction line. Friction loss in 6" pipe at 1000 gpm from table 15 of the Hydraulic Institute Engineering Data Book, is 6.17 feet per 100 feet of pipe.

in 4 feet of pipe friction loss = 4/100 x 6.17 = 0.3 feet

Friction loss coefficients (K factors) for the inlet, elbow and valve can be added together and multiplied by the velocity head:

FITTING K FROM TABLE6" Square edge inlet 0.50 32 (a) 6" 90 flanged elbow 0.29 32 (a) 6" Gate valve 0.11 32 (b)

Total coefficient, K = 0.90

Total friction loss on the suction side is:

hfs = 0.3 + 1.7 = 2.0 feet at 1000 gpm.

4. The total suction head then becomes:

hs = hss + hps - hfs = 5 + (-23.12) - 2.0 = -20.12 feet, gauge at 1000 gpm.


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1. Static discharge head = hsd = 40 feet

2. Discharge surface pressure = hpd = 0 feet gauge

3. Discharge friction head = hfd = sum of the following losses :

Friction loss in 6" pipe at 1000 gpm. from table 15, is 6.17 feet per hundred feet of pipe.


Friction loss in 6" elbow:

from table 32 (a), K = 0,29

from table 15, V2/2g = 1.92 at 1000 gpm.

Friction loss = K V2/2g = 0.29 x 1.92 = 0.6 feet

The friction loss in the sudden enlargement at the end of the discharge line is called the exit loss. In systems of this type where the area of the discharge tank is very large in comparison to the area of the discharge pipe, the loss equals V2/2g, as shown in table 32 (b).

Friction loss at exit = V2/2g = 1.9 feet

The discharge friction head is the sum of the above losses, that is:

hfd = 27.2 + 0.6 + 1.9 = 29.7 feet at 1000 gpm.

4. The total discharge head then becomes:

hd = hsd + hpd + hfd = 40 + 0 + 29.7 = 69.7 feet, gauge at 1000 gpm.

c. Total system head calculation:

H = hd - hs = 69.7 - (-20.2) = 89.9 feet at 1000 gpm.

Our next example will be the same as the one we just finished except. that there is an additional 10 feet of pipe and another 90° flanged elbow in the vertical leg. The total suction head will be the same as in the previous example. Take a look at figure # 4

Nothing has changed on the suction side of the pump so the total suction head will remain the same:

hs = -20.12 feet, gauge at 100 gpm.


1. The static discharge head "hsd" will change from 40 feet to 30 feet, since the highest liquid surface in the discharge is now only 30 feet above the pump centerline.(This value is based on the assumption that the vertical leg in the discharge tank is full of liquid and that as this liquid falls it will tend to pull the liquid up and over the loop in the pipe line. This arrangement is called a siphon leg).

2. The discharge surface pressure is unchanged:

hpd = 0 feet

3. The friction loss in the discharge pipe will be increased by the additional 10 feet of pipe and the additional elbow.


The friction loss in the additional elbow = 0.6 feet

The friction head will then increase as follows:

hfd = 29.7 + 0.6 + 0.6 = 30.9 feet at 1000 gpm.

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The total discharge head becomes:


= 30 + 0 + 30.9

= 60.9 feet, gauge at 1000 gpm.

5. Total system head calculation

H = hd - hs = 60.9 - (-20.12) = 81 feet at 1000 gpm.

For our last example we will look at gauges. Take a look at FIG 5:

Specifications:

• Capacity - 300 gpm. • Specific gravity - 1.3 • Viscosity - Similar to water • Piping - 3 inch suction, 2 inch discharge • Atmospheric pressure - 14.7 psi.

Divide the heads into two sections again:

The discharge gauge head corrected to the centerline of the pump, in feet of liquid absolute is found by adding the atmospheric pressure to the gauge reading to get absolute pressure, and then converting to absolute head:

hdg = (130 + 14.7) x 2.31 / (1.3 Specific Gravity + 4 ) = 261.1 feet, absolute


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The discharge velocity head at 300 gpm. is found in table 9 of the Hydraulic Institute Engineering Data Book

hvd = 12.8 feet at 300 gpm.

The suction gauge reading is in absolute terms so it needs only to be converted to feet of liquid, absolute.

hgs = 40 x 2.3 / 11.3 +2 = 73.08 feet absolute


The suction velocity head at 300 gpm. is found in table 11 of the Pipe Friction Manual:

hvs = 2.6 feet at 300 gpm.

The total head developed by the pump is:

H = (hgd + hvd ) - ( hgs + hvs ) = (261.1 + 12.8) - (73.08 + 2.6)= 198.22 feet absolute at 300 gpm.

Estimating the shutoff head of a centrifugal pump:

In the fifteenth century the Swiss scientist Daniel Bernoulli learned that the combination of head and velocity was a constant throughout a piping system. He then wrote the formula showing the relationship between this liquid velocity, and resultant head. As many of you know, I often quote this formula in my pump and seal schools. The formula looks like this:

• V = Velocity or speed of the liquid at the impeller outside diameter (ft/sec. or meters/sec.)

• g = gravity = 32.2 feet / second2 or 9.8 meters / second2

My students have heard me quote this formula as the basis for my statement that you can estimate the shut off head of a 1750 rpm. centrifugal pump by squaring the diameter of the impeller. How did I come to that conclusion ? Lets look at the formula again, and we will start by defining velocity:

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Velocity is a measurement of speed using distance and time as the variables. The terms we use to discuss velocity are feet/second or meters/ second. In the inch system the velocity of the impeller outside diameter is determined by the following formula:

• d = diameter of the impeller • • rpm = speed of the impeller outside diameter • 12 = twelve inches in a foot • 60 sixty seconds in a minute

Now we will solve the formula. Substituting 1750 for the rpm we would get:


This means that at 1750 rpm the shutoff head is 90% of the diameter of the impeller squared

If you will check a typical pump curve as supplied by the pump manufacturers, you will learn that the shut off head actually varies from 90% to 110% of the diameter of the impeller squared. I elected to use 100% because it is a sensible average and in some cases it accounts for the additional velocity added to the fluid as it moves from the impeller eye to the impeller outside diameter.

If we substitute 3500 rpm for the speed, the new numbers would look like this


We can round out the 3.6 to 4.0 and say that at 3500 rpm the shutoff head equals approximately the outside diameter of the impeller squared, times four.

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se decimeters instead. It will make the calculations a lot simpler because you will be

It is a little trickier in the metric system. Instead of using millimeters when measuring the impeller diameter, move over two decimal places and u

using more convenient, larger numbers.

Inserting the numbers into the formula we would get a velocity of:



If the pump were running at 2900 rpm you would get



How do we use this information? You can combine this formula with your pressure to head and come up with an estimate to

see if an operating pump is operating close to its BEP(best efficiency point ). As

ure gage reads 20 psi. The pump is pumping the difference between these readings, so the pump is pumping 100 psi.

he pressure to head conversion is:

knowledge of how to convert

an example:

In the inch system a pump discharge pressure gage reads 120 psi. The pump suction press

At its BEP(best efficiency point) the pump should be running between 80% and 85% of its shut off head. 100 psi is 83% of 120 psi. T

The pump has an 8.5 inch impeller running at 3500 rpm. The shutoff head would be (8.5 inches)2 x 4 = 288 feet. Pretty close!

In the metric system we can make the calculation for a 295 millimeter impeller

The pump discharge pressure gage reads 10 bar The pump suction pressure

ing 9 bar

turning at 2900 rpm

gage reads 1 bar The pump is pumping the difference between these readings so the pump is pump

At its BEP(best efficiency point) the pump should be running between 80% and 85% of its shutoff head. 9 bar is 83% of 10.8 bar. The pressure to head conversion is:

106 meters shut off head. The pump has a 295 mm impeller running at 2900 rpm. The shutoff head would be (2.95 decimeter)2 x12 =104.4. Pretty close!

Rules of thumb for pumps

ok at the manufacturer's published pump curve. The problem is that you do not always

mp companies test their pump to determine its performance, they have no need for general guide lines or "rules of thumb."

PUMP BASICS

p (inch sizes) o At 1750 rpm. Shut off head = Diameter of the impeller squared

3500 rpm. Shut off head = Diameter of the impeller squared x 4 o For other speeds you can use the formula : Shut Off Head = D2 x (new

ber. (6,25)

If you want to know a pumps capabilities the rules are simple, lo

have the curve available. Pu

Over the years I have accumulated many of these rules to help me estimate pump performance, here are a few of them:

• How to estimate the shut off head of a pum

o At

rpm / 1750)2 • Estimating metric head is a little bit more involved, but it still works:

o Measure the shaft in mm. ( as an example: 250 mm ) o Mark off two places. (2,5) o Square the numo For 1450 rpm, multiply by 3 (18,75) o Add 10 % for the answer in meters. (21 meters )

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would multiply by 12 instead of 3. Although se formulas you cannot estimate the

r that.

nter of the inboard bearing to the center

• modulus of elasticity changing shaft

t solution. When pump

e speed of a pump you will get twice the capacity, four times the

ft has only a small portion of the conductivity of a carbon

ack to the bearing oil.

curves for head, capacity, and

6% stock

r less e settings are determined by the pump manufacturer and

0,2 to 0,5 mm) You lose 1% of the 0,05 mm) you miss this setting.

wear rings less than two inches (50 mm.) in outside diameter.

o NOTE: For 3000 rpm, youyou can estimate shutoff head with thepump capacity. You will need the pump curves fo

• The pumps best efficiency point (B.E.P.) is between 80% and 85% of the shut off head. At this point there is little to no radial thrust on the impeller. Also the "power in" is closest to the "power out".

• The L3/D4 ratio should be below 60 (2.0 in metric) to prevent excessive shaft bending. To calculate it for end suction centrifugal pumps :

o L = length of the shaft from the ceof the impeller (inches or millimeters). Caution: do not use centimeters, the numbers will come out wrong.

o D = diameter of the shaft (under the sleeve) in the stuffing box area (inches or millimeters) Do not use centimeters.

Since most shaft materials have a similar materials will not prevent shaft bending when you operate off of the B.E.P. Lowering the L3/D4 is the only logical and efficienmanufacturers discuss operating off of the B.E.P. they relate problems to the heat that will build up in a minimum flow condition and ignore the problems with shaft bending.

• A double suction pump can run with 27% less N.P.S.H. or at a 40% faster speed without cavitating.

• If you double thhead and it will take eight times the horsepower to do it.

• A stainless steel shasteel shaft. This is very important when you are pumping at elevated temperatures because we do not want to transmit the high temperature b

• If you double the speed of a pump you will get almost four times the shaft whip, wobble or run out and eight times the wear.

• Multistage pumps reduce efficiency 2% to 4%. • In many instances an inducer can lower Net Positive Suction Head Required by as

much as 50% . • If you are pumping paper stock, modify the

efficiency as follows: o 0.725 foro 0.825 for 5.5% stock o 0.90 for 5% o 0.94 for 4.5% o 0.98 for 4% o 1.0 for 3.5% o

• Open impeller clearancnormally run between 0.008" and 0.015" (pumps capacity for each 0.002" (

• Wear ring clearances are very similar to impeller clearances, but you lose 1% pump capacity for each 0.001" (0,025 mm) of wear. A typical clearance would be 0.003 inch/inch diameter with 0.010 inches (0,3 mm) minimum clearance for

119

face seals in these

tive face

e bearing race of a properly installed

F) temperature increase. This corresponds to :

Th il is not being contaminated by wa

• Packing leakage

earing case.

running at 1750 rpm. would cover about 100,000 miles ning

tomatic transmission oil every 25,000 miles ( 40,000 kilometers)

ping temperature exceeds 200 degrees Fahrenheit (100° C). This design will allow the wet end of the pump to expand in

ions instead of from the feet up, destroying the wear rings.. • Try to buy pumps with a Suction Specific Speed (SSS) below 8,500 (10,000

• Bearing grease or lip seals have a design life of less than 2000 hours. In a constantly running pump this would be only 83 days. These seals will also damage the expensive shaft and place a stress point at the maximum bending moment arm. Substitute non fretting labyrinth seals, or positivelocations. It is a good idea to install them in electric motors also to prevent moisture from entering and damaging the motor windings and bearings.

• Do not use a vent on the top of the bearing case. At shut down the outside moisture will enter the bearing housing through this vent. Let the moisture attempt to enter the case through the labyrinth seals instead, they will do a better job of directing the moisture to the external drain hole. If you install posiseals you can forget about this problem.

• The axial clearance in a bearing is ten times the radial clearance. This is the reason proper installation is so critical. If the bearing is over compressed the bearing balls will distort and roll instead of spin causing excessive heat and premature failure. The temperature at thbearing is at least 10 degrees Fahrenheit (5° C) higher than the oil sump temperature.

• The life of bearing oil is directly related to its temperature. The rule of thumb used by the SKF Bearing Company is that the service life of an oil is specified as 30 years at 30 degrees Centigrade (86° F) and is cut in half for each 10 degree Centigrade (10

o A life of 3 months at 100 C. (212 F.) o A life of 6 months at 90 C. (195 F.) o A life of 12 months at 80 C. ( 176 F.)

ese numbers assume that the lubricating oter from one or all of the following sources:

• The water hose used to wash the packing leakage away from the pump area. • Aspiration, as moisture laden air enters the b

An automobile engine (160,000 kilometers) every 2000 hours (83 days in the life of a constantly runpump ). Auto manufacturers recommend changing their au

APPLICATION

• Use Centerline pump designs when the pum

two direct

metric) Do not buy pumps with a SSS over 12,000 ( metric 16,500) unless you are pumping hot water or mixed hydrocarbons. If you have a double suction pump you can divide the SSS number by 2

120

. If you have a high static or pressure

e best efficiency point.

ly .

d on long shaft vertical pumps to prevent

ad greater than 650 feet (198 meters) and more

ency vibrations and low frequency vibrations at reduced flow

hould be at least 10 diameters of pipe between the suction of the pump and the first elbow. This is especially critical in double ended pump designs as the

inlet flow can cause shaft thrusting, and subsequent bearing problems. • Substituting a globe valve for a gate valve in a piping system is similar to adding

in the

speed cuts in. You will

side down.

• Do not specify a pump with the largest impeller available . Give yourself an additional 5% or 10% you might need it.

• The maximum viscosity a centrifugal pump can handle would be a product similar to 30 weight oil at room temperature.

• Use a variable speed pump if your head is mainly system head. Circulating hot or cold water would be typical applicationshead, as is the case with a boiler feed pump, the variable speed will not be of much help in keeping you on or near th

• Pumps piped in series must have the same capacity (impeller width and speed) • Pumps piped in parallel must have the same head (impeller diameter and speed ) • Use a rotary positive displacement pump if your capacity is going to be less than

20 gpm.(4,5 cubic meters per hour) • A centrifugal pump can handle 0.5% air by volume. At 6% it will probab

become air bound and stop pumping. Cavitation can occur with any amount of air• Use double volute pumps any time your impeller diameter is 14 inches (355 mm)

or greater. They should also be useexcessive shaft movement that will cause problems with the packing, seals, bearings and critical dimensions.

• A Vortex pump is 10% to 15% less efficient than a comparable size end suction centrifugal pump.

• The A.P.I. (American Petroleum Institute). sixth edition states : High energy pumps, defined as pumping to a hethan 300 horsepower (224 KW) per stage, require special consideration to avoid blade passing frequrates.

PIPING ETC..

• There s

turbulent

another 100 feet (31 meters) of piping to the system. On the discharge side of the pump this will cause the pump to run off of its B.E.P. with a resultant shaft bending. On the suction side of the pump it will probably cause Cavitation.

• After the pump and motor have been aligned, dowel both the pump and the motor to the base plate. Be sure to dowel only the feet closest to the coupling, allowing the outboard ends to expand with temperature changes.

• Check impeller rotation after installing the pump. Do not assume it will turn correct direction. I have heard about two speed pumps with the second speed wired backwards. They will drive you crazy because the pump will often meet its head requirement but not the capacity when the secondalso notice excessive noise at this time.

• Use eccentric reducers rather than concentric reducers at the pump suction. Concentric reducers will trap air. Be sure the eccentric reducer is not installed up

121

d level falling greater than 3 Ft./sec. (1 Meter/ sec.)

t velocities in pipes leaving vessels. Generally greater than 10

apor point. ditions.

onsult the

t be at least one and one half diameters of

TROU

low. o Air is entering at the pump suction.

uid turbulence at the pump suction. • Cavitation damage on the leading edge of the impeller blade indicates internal

below 9000 hat the problem is with the impeller ted when the pump manufacture tried

, as a result, often experience this

of 6 inches (150 mm.) wider.


• Vortexing can occur if any of the following conditions are present: o Low liquid levels o Liquio There is a large concentration of dissolved gases in the liquid. o High outle

feet/sec. (3 meters/sec.) o Liquids near their vo High circulation caused by asymmetrical inlet or outlet cono Inlet piping too close to the wall or bottom of the tank. C

Hydraulic Institute Manual or a similar publication for recommended clearances.

o In a mixer, the liquid level musthe blade, above the blade.

BLESHOOTING

• Cavitation damage on the trailing edge of the impeller blade means : o The N.P.S.H. available is too

o There is liq

recirculation. Check the Suction Specific Speed number to see if it is (10,000 metric). Higher numbers mean tshape or adjustment. The problem was creato come up with too low a N.P.S.H. Required.

• Cavitation damage just beyond the cutwater, on the casing and tip of the impeller blade, indicates the impeller blade is too close to the cutwater. This clearance should be at least 4% of the impeller diameter up to a 14 inch (356 mm.) impeller, and 6% greater than 14 inch ( 356 mm.). Some self priming pump manufacturers want a maximum clearance of 1/8" (3 mm) andproblem. A repaired or substituted impeller is often the cause of the problem in a non self priming pump.

• Water in the bearing oil will reduce bearing life 48%. The water enters from packing leakage, wash down hoses, and aspiration caused by the temperature cooling down in the bearing casing after shutdown and moisture laden air entering the bearing case. A 6% water content in the oil will reduce bearing life by as much as 83%

• The mass of the pump concrete foundation must be 5 times the mass of the pump, base plate, and other equipment that is being supported, or vibration will occur.

• Up to 500 horsepower (375 KW), the foundation must be 3 inches (76 mm.) wider than the base plate all around. Above 500 horsepower (375 KW) the foundation should be a minimum

122

sitive face sealing at the bearings, if you could solve the emission problem.

the

ir will enter the

eater drain lication.

the suction

rloading power will occur if the pump is run

Horsepower is measured using the units, foot pounds with one horsepower equal ounds

per minute we are pumping by finding out how much a gallon of our fluid weighs.

the pump is producing and you have foot pounds per minute that can be converted to horsepower.

• Imaginary lines extended downward 30 degrees to either side of a vertical through the pump shaft, should pass through the bottom of the foundation and not the sides.

• The bearing oil level should be at the center of the lowest most ball of a stationary bearing. The preferred choice for bearing lubrication would be an oil mist system with po


The top of the reducer should go straight into the suction flange. • Valve stems, T Branches and elbows should be installed perpendicular to

pump shaft, not at right angles. • Do not use packing in any pump that runs under a vacuum, as a

system through the pump stuffing box.. These applications include : o Pumps that lift liquid. o Pumps that take their suction from a condenser or evaporator. o Any pump that takes its suction from a negative pressure. H

pumps are a typical app• Be sure too vent the stuffing box of a sealed, vertical pump back to

side of the pump or air will become trapped in the stuffing box. The vent must be located above the lapped seal faces.

• If the Specific Gravity of the pumping liquid should increase, due to temperature, there is a danger of overloading the motor and therefore motors having sufficient power should be used. The same ovetoo far to the right of its B.E.P.. This is a very common problem because of the great number of oversized pumps in existence.

How to the calculate the water horsepower coming out of the pump?

to 33,000 foot pounds. Since fluid has weight we can calculate how many p

After you have done that, multiply the gallons per minute you are pumping by 8.33 (the weight of a gallon of water) and then multiply that result by the specific gravity (the weight) of your fluid, and you will have the pounds per minute number you are looking for.

Once you have the pounds per minute you are pumping, you can multiply that number into the feet of head

Please take a look at the following pump curve.

Let's use this chart for our example:

You are using the 13 inch impeller at 1750 rpm and pumping 300 gallons per minute of a fluid with a specific gravity of 1.0 to a head of 168 feet

300 gpm x 8.33 x 1.0 sg. = 2,499 pounds of fluid per minute.

We are pumping this fluid to a head of 168 feet so:

2,499 x 168 = 419,832 foot pounds per minute

Since 33,000 foot pound per minute equals one horsepower. We will divide and get:

419,832 / 33,000 = 12.73 horsepower.

This means that the pump is putting out 12.73 horsepower. Now, the next question is how much actual horsepower is required to do this?

Please take a look at the ascending lines on the bottom of the chart. Each line represents a different size impeller with the top line showing the horsepower required for a 13-inch impeller and the bottom line for a 9-inch impeller. The horsepower required is shown in the left column under bhp. (brake horsepower).

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Notice that it calls for 20 horsepower to move 300 gallons per minute with a 13-inch impeller.

If the pump were 100% efficient, all you would need would be 12.73 horsepower motor to drive the pump and it would do the job, but motors and pumps are not 100 % efficient because of friction losses and heat generation. This means that our actual efficiency is

12.73 hp out / 20 hp in = 0.64 or 64% efficient

Suppose the specific gravity of the fluid you are pumping is different than 1.0 (cold water). Just plug the new number into the formula and multiply the pump curve bhp by the same number. Using a specific gravity of 1.3, the change would look like this:

300 x 8.33 x 1.3 = 3248.70 pounds per minute

3248.70 x 168 = 545,781.60 foot-pounds per minute

545,781.6 / 33000 = 16.54 water horsepower out of the pump

20 x 1.3 = 26 horsepower is going into the pump

16.54 / 26 = 0.64 efficient

A few things you should know about your pump's piping system

• There should be at least 10 diameters of pipe between the suction of the pump and the first elbow. This is especially critical in double-ended pump designs as the turbulent inlet flow can cause shaft thrusting, and subsequent bearing problems. If an elbow must be installed be sure it is in a plane at right angles to the pump shaft to prevent an uneven flow to both sides of a double suction impeller.


The top of the reducer should go straight into the suction flange. • Piping should be arranged with as few bends as possible. If bends are necessary

use a long radius when ever possible • Valve stems, T Branches and elbows should be installed perpendicular to the

pump shaft, not at right angles. • If an expansion joint is installed in the piping between the pump and the nearest

point of anchor in the piping, It should be noted that a force equal to the area of the expansion joint (which could be a lot larger than the normal piping size) times

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125

the pressure in the piping will be transmitted to the pump proper. Pipe couplings that do not provide an axially rigid connection have the same affect. If an expansion join or non-rigid coupling must be used, it is recommended that a pipe anchor be installed between it and the pump.

• Be aware that radial forces are being generated in the pump housing from the pressure in the piping system acting on the volute area. The magnitude and direction of the forces is dependent upon the piping arrangement along with the areas and pressures involved.

• It is always a good idea to increase the size of the suction and discharge pipes at the pump nozzle in order to decrease the head loss from pipe friction.


• If increasers are used on the discharge side to increase the size of discharge piping, they should be installed between the check valve and the pump.

• Both a check and gate valve should be installed in the discharge piping with the check valve placed between the pump and the stop valve to protect the pump from reverse flow and excessive back pressure. Manually operated discharge valves that are hard to reach should have some facility for quick closing. A sprocket rim wheel and chain or a remotely operated motor are two alternatives you might consider.

• Suction piping must be kept free of air leaks. • The installation of check valves should be avoided in the suction piping although

they are often used to reduce the number of valves that have to be operated in switching between series and parallel pump operation.

• A foot valve is often installed in the suction piping to aid priming. Do not install them if the pump is operating against a high static head because failure of the driver would allow liquid to rush back suddenly causing water hammer. This is especially true for vertical turbine and submersible pumps that are not designed for use with a foot valve.

• Foot valves should be of the low loss flap type rather than the multiple spring variety and have a clear passage for the liquid at least the same area as the suction piping.

• A horizontal suction line should have a gradual rise or slope to the pump suction. • Cast iron pumps should never be provided with raised face flanges. If steel

suction or discharge piping is used, the pipe flanges should be of the flat face type and not the raised face type. Full-faced gaskets must be used with cast iron flanges.

• The optimum control valve location is within five feet (1,5 meters) of the pump discharge to prevent too much surging of fluid in the system when the discharge is throttled.

• The optimum pipe size will consider the installed cost of the pipe (the cost increases with size) and the pump power requirements (the power required increases with pipe friction)

o Try to limit the friction loss at design flow to 2-5 feet for each 100 feet (1-2 meters for each 30 meters) of pipe).

o To prevent the settling of solids you need a minimum velocity of about 4 to 7 feet per second (1.5 to 2.5 meters per second)

o Velocities of no more than 10 feet (3 meters) per second are recommended in the suction side piping to prevent abrasive wear.

A few rules of thumb for mechanical seals

Before selecting your mechanical seal design there are three things you want to remember:

• All of the seal materials must be chemically compatible with any fluids that will be pumped through the system and that includes solvents, cleaners or steam that might be introduced into the system to flush or clean the lines. It also includes any barrier fluids that are used to circulate between dual mechanical seals.

• The seal faces must stay together. If they open the seal will leak and allow solids to penetrate between the faces where the solids will eventually destroy the lapped surfaces.

• Good seal life is defined as running the mechanical seal until the carbon face is worn away. Any other condition is called a seal failure and is always correctable

The following is offered as a guide when dealing with mechanical seals in general. If possible you should contact the manufacturer for specific recommendations and limits. I have spent the past twenty seven years lecturing about seals and pumps and during that time have picked up a number of rules that are worth remembering. Here are some of the most important:

• Selecting materials - The elastomer ( the rubber part) • There are two temperature limits for a mechanical seal:

o You must not exceed the temperature of the seal components. As an example Ethylene Propylene rubber cannot seal hot fluids in excess of 300° degrees Fahrenheit ( 150° C) without taking a compression set and eventually leaking.

o You must not exceed the temperature limit of the fluid you are pumping. Many fluids will change from a liquid to a gas, solid or crystal at elevated temperature. In almost every case this will cause a seal failure. As an example, petroleum lubricating oil cokes between 250 and 300 degrees Fahrenheit (120° C. to 150° C.) and restricts the movement of the seal components. A Viton® O-ring, in this application would not have been subjected to its temperature limit, but we had the seal failure because we exceeded the temperature limit of petroleum products.

• Halogens will attack Teflon® coated elastomers . Halogens are easily identified because they end in the letters " INE". The list would include Bromine, Chlorine. Astatine, Fluorine, and Iodine. These Halogens will penetrate the Teflon®

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coating and attack the base rubber material causing it to swell and split the Teflon sleeve or coating.

• Most Viton® compounds are attacked by water. Be sure to check if you have the correct one. Remember that steam is another name for water and the steam cleaning of lines is very common in the process industry. Caustic is another common cleaner and caustic contains a high percentage of water also.

• Buna "N" (Nitrile) is an elastomer that has a short shelf life. This is the elastomer that is most often used in Rubber Bellows Seals. The problem is Ozone attack. Ozone is produced by the sparking from electric motors, so it is a very common problem. A typical shelf life for most Buna compounds would be one year.

• If a round O-Ring becomes square in operation (compression set) it is almost always caused by excessive heat. Chemical attack is usually recognized by a swollen and soft elastomer while high heat will produce a shrunken, hard one.

• Chemical attack of the elastomer will usually cause a seal failure within five to ten days. The swollen elastomer will "lock up" the mechanical seal and in some instances, open the lapped seal faces.

Determine the correct O-Ring by one of the following methods:

• Look up the chemical in published O-Ring charts provided by all reputable seal companies. You will find a chart in the chart section of this web site

• Check to see if the plant has any experience with O-Rings, in this fluid, in another seal application. O-Rings can also be found in filters, strainers, valves, flanges, expansion joints etc..

• Test the O-Ring by immersing it into the sealing fluid for one week. If the O-Ring changes weight, shape, or appearance, it is not compatible with the fluid.

• Use a universal O-Ring compound such as Green Tweed's Chemraz, Dupont's Kalrez® or a similar product.

• When choosing an O-Ring, or any other elastomer, be sure to consider any cleaners or solvents that might be flushed through the lines or that could come into contact with the seal. The elastomer must be compatible with these fluids also.

• Never use " glued together" elastomers in a split seal or any "dynamic" application. A hard spot will be created that will interfere with the movement of the dynamic elastomer.

Selecting Materials - The Faces.

• Carbon and most hard face materials have an expansion rate of about one third that of stainless steel.

• Use two hard faces if the product has a tendency to solidify between the seal faces. Never use plated or coated hard faces in these applications. Hard faces are recommended if you find that it is impossible to keep the seal faces together and solids are present in the sealing liquid. Two hard faces are also recommended in the sealing of hydrocarbons that have to pass a "fugitive emissions" test. Coke

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particles forming between the faces will pull pieces of carbon out of the carbon/graphite face presenting a leak path for fugitive emissions.

• Although many carbon graphite compounds are available unfilled carbons are the best because they are corrosion resistant to almost all chemicals except oxidizing agents and some de ionized water applications. These oxidizing agents will combine with the carbon to form Carbon Monoxide and Carbon Dioxide. The most common oxidizers are oleum, sulfur trioxide, strong bleaches and nitric Acid. You cannot use any form of carbon in these applications. Keep in mind that black elastomers will also be attacked by oxidizing agents because of their carbon content.

• Ceramic vs. ceramic is a good choice for oxidizing chemicals. • If you are going to select plated Tungsten Carbide as a face material, use only the

nickel base Tungsten Carbide. Cobalt base is too hard and can crack with normal seal face differential temperatures. Nickel base, because of its superior corrosion resistance is the preferred material for solid Tungsten Carbide faces also.

• Reaction bonded Silicone Carbide has excellent wear characteristics, but contains up to 17% free silica which can be attacked by many chemicals including caustic. Alpha sintered Silicone Carbide is also available and is Silica free.

• 85% ceramic should never be recommended as a hard seal face as it can break with as little as a 100 degree Fahrenheit (55 C) temperature difference. 99.5% would be a much better choice.

• Plating or coating a seal face will not give it corrosion resistance. Coatings are used for wear resistance and low friction. To get corrosion resistance the outer coating must be at least 1/8" (3 mm) thick. If the base material is not corrosion resistant to the pumping fluid and any cleaners or solvents used in the lines the corrosive will go through the coating and attack the base, causing the plating to come off in sheets.

Selecting Materials - The Metal Parts.

• Be sure to use low expansion metal such as Carpenter 42 or Invar 36 in your metal bellows seal face holder if the product temperature can exceed 400° Fahrenheit (205°C). These low expansion steels will prevent the carbon or hard seal faces from leaking between the face and the metal holder. Needless to say glue or epoxy is not a sensible solution to differential expansion problems.

• If your pump is manufactured from Iron, steel, stainless steel, or bronze, you can probably use a seal manufactured from 316 stainless steel components. The springs or bellows, however, must be manufactured from Hastelloy "C" to avoid problems with Chloride Stress Corrosion.

Sealing Limits

• Use only stationary mechanical seals (the springs do not rotate with the shaft) if the face surface speed exceeds 5000 feet per minute ( 25 M/sec.), but never in a cartridge design unless some method has been provided to insure that the cartridge sleeve is square to the shaft.

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• Use O-Ring balanced seals in vacuum applications down to 10-2 inches or one millimeter of mercury (1 Torr.). The O-Ring is the only elastomer that can seal both vacuum and pressure. Split seals will work in these applications, but they must be turned around for best operation.

• Any good quality, balanced, O-Ring seal can seal stuffing box pressures to 400 psi (28 bar) and temperatures to 400 degrees Fahrenheit (205° C). There is a compound of Dupont's Kalrez® that is satisfactory to 600 degrees Fahrenheit (370° C), but it is not acceptable at ambient temperatures (it gets too hard).

Application

• A Balanced O-Ring seal will not vaporize the product at the seal face if the stuffing box pressure is at least one atmosphere above the products vapor point.

• The easiest product to seal is a cool, clean, lubricating liquid. All problem chemicals can be placed into several categories. If you know how to seal these categories you should have no trouble making seals work in your applications :

o Products that crystallize (caustic or sugar solutions) o Viscous products (asphalt or molasses) o Products that solidify (polymers or chocolate) o Products that vaporize (hot water or benzene) o Film building liquids (hot petroleum or plating solutions) o High temperature fluids (heat transfer oil or liquid sulfur) o Dangerous products (fire hazard, explosive, radioactive, bacteria) o Non lubricating liquids (solvents or hot water) o Gases and dry running applications (hydrogen) o Dry solids (cake mix or pharmaceuticals) o Corrosive fluids (acids or strong bases) o Cryogenics (liquid nitrogen) o Slurries (river water, sewage, most raw products)

• In addition to these chemical categories there are other sealing problems that include:

o High pressure o Hard vacuum o High speed o Excessive motion

• Dual seals should be balanced in both directions to prevent failure when barrier fluid pressure changes. The practice of using "one direction" seal balance is commonly employed by most seal companies and should be avoided for both safety and reliability.

• Use motion seals on mixers, agitators, sleeve bearing equipment and any rotating device that has motion greater than 0.005" (0,15 mm.) in a radial or axial direction. Pump seals do not work well in these applications because the hard faces are too narrow and the internal seal clearances are too tight.

• Do not use flushing fluid as a coolant in stationary mechanical seals. The coolant will be directed to only one side of the seal and since a stationary seal does not rotate the sliding components the differential temperature can cause the faces to

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go out of flat. In the case of stationary bellows seals it could cause a bellows rupture.

• The best way to cool a seal is to use the jacketed stuffing box that came as a part of the pump. This jacket will not only cool down the seal area, but will provide the necessary cooling to the shaft so that it will not transmit stuffing box heat back to the bearings.

• The use of steam in a Quench gland is another solution, but not as good as the jacketed stuffing box.

• It is all right to dead end fluid in a stuffing box if a jacketed stuffing box is being used. Do not attempt to recirculate back to the suction side and cool the stuffing box at the same time. When using a jacketed stuffing box it is best to install a carbon bushing in the bottom to act as a thermal barrier the pumping fluid and the seal.

• Do not use rotating, "Back to Back" double seals in dirt or slurry service. The solids will prevent the inner seal from moving forward as the faces wear and if the barrier fluid pressure is lost, solids will penetrate the inner seal faces.

• Be sure to vent vertical pumps back to the suction side of the pump. Air trapped in the stuffing box can cause the seal faces to run hot and in some instances destroy the elastomer.

• Cyclone type separators or "in line filters" are not a good method of cleaning up the fluid in the stuffing box.

• Heat affects a seal several ways: o The faces can be attacked. Plated faces can have the hard coating crack off

and filled carbons can have the binder melted out in high heat. o The elastomer (rubber part) has a temperature limit determined by the

compound used. o The corrosion rate of all liquids increases with temperature. o Thermal expansion can cause seal face loads to alter and seal face flatness

to change. o Many products will change from a liquid to a solid or gas in the presence

of high temperature. If this should occur between the seal faces, they can be blown open.

• Do not be tempted to put the mechanical seal outside of the stuffing box to keep the springs out of the fluid. As the face wears the seal must move into the slurry where it will eventually "hang up" and leak. In these applications centrifugal force is throwing solids into the lapped faces and if there is excessive pressure in the system the seal faces will be blown open.

• When choosing the pressure range of a mechanical seal be sure to consider the stuffing box pressure not the pump discharge pressure. Very few seals will ever see discharge pressure.

Technical

• Seals lapped to less than three helium light bands ( 0.000034") inches or 1,0 microns) should not show visible leakage. Visible leakage occurs at about 5 light bands.

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• A typical mechanical seal face load would be 30 psi. (0,2 N/mm2) when the carbon is new and 10 psi. (0,07 N/mm2) when the carbon is fully worn away. You must never guess as to how much to compress a mechanical seal. Either take the information from the seal print or calculate the correct length from the above information.

• Both rotating and stationary metal bellows seals require vibration damping. Elastomer seals do not experience this vibration problem because the elastomer touching the shaft is a natural vibration damper. Vibration can be either harmonic or caused by poor lubricating fluids (slip stick)

• Use only non fretting seal designs. Shafts and sleeves cost too much to ignore this severe problem.

• Carbon throat bushings should have a shaft clearance of 0.002 inches/inch (0,002 mm/ millimeter) of shaft diameter. If they are to be used as a support bearing you should cut the clearance down to 0.001 inches/ inch (0,001 mm/millimeter) of shaft diameter.

• It is not necessary to lubricate seal faces at installation. If the product you are sealing can vaporize between the faces and cause freezing then you must remove any lubricant that might have been placed there by the manufacturer.

• Balanced mechanical seals consume about one sixth the horsepower of packing. Packing a pump would be like running your automobile with the emergency brake engaged. The car would run, but the fuel consumption would be high.

• Single spring seals are wound in either a right or left handed direction. Check to see if your seal has a problem in keeping the faces together because of the spring winding.

• Open impeller pumps require impeller adjustment. Use only cartridge or split seals in these applications. Do not use seals that locate against a shoulder or set screw to the shaft, as the face load will change when the impeller is adjusted.

• Do not relap the carbon face unless it is an emergency. Seal face opening is a common seal failure. When the faces open solid particles imbed them selves into the carbon face and will be driven in even further during the lapping process. If you must relap in an emergency never use lapping powder, as the abrasive particles will imbed into the soft carbon.

• You cannot balance an inside seal by removing material from the carbon face. To get seal balance you must do one of the following:

o Use a stepped sleeve with rotating seals. o Let the carbon slide in a case that is sealed to the shaft. o Use a metal bellows. The balance is not perfect, but good enough. o Use a stationary seal design, they require no stepped sleeves.

• Seal face hardness is a confusing subject because of the various measuring scales employed. The two most common are Rockwell "C" and Brinnell. If you divide the Brinnell scale by ten (10) it is almost equal to the Rockwell "C" scale.

• Avoid oil as a barrier or buffer fluid between two mechanical seals. Most petroleum base and other oils have a low specific heat (0.2 - 0.4) and combined with poor conductivity (0.5 of water) makes them a poor choice compared to fresh water. If oil is mandatory, a clean heat transfer oil would be your best choice.

• A convection tank can often be used between two balanced O-Ring seals. If you use unbalanced seals the heat generated by this type of seal is usually excessive for convection cooling. Contact the seal manufacturer for his recommendations concerning speed, diameter, face combination and pressure limits for convection cooling. If convection is not satisfactory, a pumping ring or forced lubrication is another option.

• If you decide to repair your mechanical seals in house, be sure to purchase the parts from the original manufacturer. If you decide to have them repaired send them back to the original manufacturer. It is important that the seal be rebuilt with the original materials and it must meet the original tolerances. This information is not available from the manufacturer because of product liability problems.

• O-ring seal designs can tolerate three to four times the "run out" capability of sliding or pusher seals incorporating wedges, chevrons, U- cups etc..

• Oil on the seal faces can cause the faces to stick together during long periods of non running. If you do not intend to run the equipment soon remove any oil that might be on the seal faces during the assembly procedure.

A quick reference to prevent potential seal and pump problems:

The biggest advantage of experience is you have hopefully learned what can get you into trouble. The following information has been explained in detail in previous technical papers, but I still see the same problems re-occurring on a daily basis.

Take a few minutes and look at the following. It might save you a seal or pump failure.

MATERIALS

• Carbon seal face. Any form of carbon is usually not acceptable in the following applications:

o Oxidizers, they combine with carbon to form CO & CO2 o Halogens (most of them end in the letters "ine") chlorine, bromine,

fluorine, astintine & iodine o Where color contamination can be a problem. o Some de-ionized water applications. o Hot petroleum products if you are concerned about fugitive emissions.

• A special carbon is used for cryogenic and hot dry air applications. Moisture is needed to make the graphite release from the carbon-graphite mixture, and in these applications the needed moisture is not present. A special carbon with an imbedded organic is made to satisfy these applications.

• Ceramic grade 99.5 is not a satisfactory hard face in hot applications because of its poor thermal conductivity. Alpha grade silicone carbide or tungsten carbide are much better choices.

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• Ethylene Propylene Rubber O-Rings will be attacked by petroleum products and this includes any petroleum grease that might be put on the O-ring during the installation process.

• Kalrez® grade 3018 is not satisfactory if the temperature is below 600°F. (315°C.) The material is too hard at these lower temperatures.

• Nickel base tungsten carbide can cause galvanic corrosion problems with stainless steel shafts.

• Reaction bonded silicone carbide is not satisfactory for caustic or most high PH materials.

• Viton® O-rings are not generally satisfactory in water based fluids. This also includes steam cleaning or flushing the lines with water based caustic solutions. Grade 747-75 fluorocarbon is O.K. if the water is cold, but ethylene propylene rubber is still your best choice as long as the temperature does not exceed 300°F (150°C.).

• White Chemraz is not recommended for most high PH fluids. Do not use it with: o Acetaldehyde, Ammonia + Lithium metal solution, Aqua Regia, Black

liquor, DI water, Ethyl Formate, Ethylene Oxide, FC 75, Freon 113 -114 - 114B2 - 115 - 142B- C318 - PCA - TF, Fuming Sulfuric Acid, Green Sulfate Liquor, KEL-F- Liquids, Lye, Magnesium Hydroxide, Red Fuming Nitric Acid, Potassium Hydroxide, Sodium Hydroxide (Caustic), Fuming Sulfuric Acid, and White Liquor.

• If you choose the wrong elastomer it will be attacked by the fluid and break down. For the first few days the seal will work very well because the elastomer has become "slimy" and moves easily. The elastomer will then "swell-up" and lock-up the moveable seal components.

APPLICATION

Remember that chemical attack can be accelerated by temperature, fluid concentration, and stress. Past plant experience is your best indicator of what seal and pump materials to use.

• Ammonia compressor; use Neoprene for the O-ring because the fluid is a combination of ammonia and petroleum oil.

• Black Liquor, as found in paper mill applications can be either sulfite or sulfate. Sulfate (high PH) is the most common and ethylene propylene can be used for the O-ring material if the temperature is below 300°F (150°C). If the temperature is too high, Kalrez is a good choice. White Chemraz is not recommended in these higher temperature caustic applications.

• Boiler feed pump applications vary a great deal. In some cases they are nothing more than a simple hot water application, but in other instances a very high pressure is involved. In any case, cooling is needed in the stuffing box to insure long seal life. High pressure applications also require a heavy duty seal design.

• Caustic. If the concentration is over 50% Monel metal will probably be needed. The metal selection depends upon the temperature and stress.

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• Ethylene Oxide will penetrate into most elastomers and explode out the other side of the O-ring. Use two seals and pressurize between them. Ethylene Oxide is a dangerous product, so two seals should be used in any case.

• Halogens attack most carbon faces and will penetrate the Teflon® encapsulated O-rings like Vanway, Creavey and & 76 style.

• Hot oils. Coking is always the problem. The seal area must be cooled. Coking is a function of temperature and time and is independent of the presence of oxygen. If you want to seal fugitive emissions you will have to go to two hard faces. Even the best of carbons show some blistering in these applications. In other words, a metal bellows seal will not eliminate the need for stuffing box cooling.

• High temperature applications. Most metal bellows seal designs incorporate a low expansion holder (Invar 36 or Carpenter 42) to retain the carbon face. This holder is also frequently used as a vibration damper to prevent seal face separation problems caused by "slip stick." If you lose cooling in these applications the pump shaft expands at a rate three times that of the low expansion steel vibration damper and can cause the seal faces to be pulled open.

• Kaoline (china clay) will penetrate lapped seal faces because the solids are less than one micron in size. You will need two seals with a pressurized barrier fluid between the seals. Water is a good choice for this barrier fluid.

• Latex balls up between the seal faces. Dual seals with a pressurized water barrier fluid have been used in this application, and non contacting gas seal seem to be the current choice, but flushing with a small amount of cold water seems to be the only satisfactory solution to this application.

• Paper stock always requires a small amount of flushing water. You cannot use suction recirculation and centrifugal force to separate the stock from the water because of the stock's low specific gravity. If the pump is trying to "lift" paper stock it will almost always cavitate.

• Pipe line applications almost always involve high pressure. Heavy duty seals should be used in these applications.

• Products that freeze (cryogenic). Watch out for moisture outboard of the seal. Dual seals with anti-freeze circulating in a convection tank is your best bet. Do not put any grease on the seal faces. It will freeze also.

• Salt water. Coat the O-rings and all clamped surfaces with Zinc Oxide paste to prevent corrosion at these locations.

• Sulfuric acid. Alloy 20 metal is usually needed for these applications. Any leakage will cause severe corrosion as the product is diluted.

CONVERTING FROM PACKING TO SEALS

Horizontally split pumps:

• Suction recirculation will not work if the stuffing box is at suction pressure. Most single stage designs fit into this category

• The face of the stuffing box must be resurfaced to get a good gasket seal.

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• If you are making a new gasket between the casing halves, be sure to have it extend outside the stuffing box face and then trim it flush after the halves are tightened together.

• Be sure to seal between the sleeve and the impeller. This is a potential leak path after a mechanical seal is installed.

• Some sleeves terminate under the seal. Check that you will not have a corrosion problem if the sleeve and shaft are different materials,

• Sometimes a new gasket will extrude into the sides of the stuffing box when the two halves of the pump are bolted together. The gasket can then rub against the side of the seal interfering with its movement.

• You will need either a stationary mechanical seal or some type of self aligning feature to seal these pumps successfully.

Flyte sewage pumps can be converted to a single mechanical seal if a special adapter is made. It's worth the problem. You only have to seal the bearing cavity in this application

MISCELLANEOUS

• Discharge recirculation can act as a sand blaster against the seal body. This can be a big problem with the thin metal plates found in metal bellows seals.

• Dual seal barrier or buffer fluid. Oils should be your last choice as a barrier or buffer fluid because of oils' low specific heat and poor thermal conductivity. You will definitely need a pumping ring if you are going to use a convection tank.

• Quenching. An excess of water or steam can easily get into, and ruin the bearings. • Suction recirculation is not affective in the following:

o Duriron pumps, because of their semi- open impeller design. o If the fluid is close to its vapor point. flashing will occur when the stuffing

box pressure drops. o If the specific gravity of the solids is lower than the fluid. If the solids

float, centrifugal force will throw the liquid to the outside leaving the solids against the seal components. Paper stock is a good example of this.

o Single stage, double ended pumps where the stuffing boxes are at suction pressure.

• Seal set screws are normally manufactured from corrosion resistant materials and are therefore softer than normal set screws. This means they can slip if reused. You can substitute hardened set screws in most cartridge seal applications.

• Do not use any type of set screw on non-metallic shafts. Seals must be clamped to the non metallic shaft or sleeve.

• Split seal designs. Most can seal either a pressure or a vacuum, but not if the application alternates between them. You can run into this problem in some mixer applications.

• Troubleshooting hints o Are other seals working in this application? If they are, you know the

materials are alright. Now you must decide what is different about this application.

o Has the seal been repaired? You may be looking at a rub mark, discoloration, or corrosion that is not relevant to this application.

CENTRIFUGAL PUMPS

• Do not let the welder use the pump as an electrical ground. You can ruin the seal or bearings in the process.

• Pumping off of the best efficiency point will not excessively deflect the shaft with the following centrifugal pump designs:

o Double volute casings. o Multi stage designs. o Diffuser or turbine pump designs.

• Be sure to level the pump when you do an alignment. • If you trim the impeller, file the tips and re balance the assembly. • The next time that you look at the pump discharge gauge, remember that the

pump pumps the difference between the suction and discharge heads. You must subtract a positive suction head to determine what head the pump is really creating.

• Bearing lip or grease seals have a useful life of less than 90 days and will cut and score the shaft because of fretting corrosion.

• Never cool a bearing housing because it will shrink and over compress the bearing. Cool only the bearing oil.

• Flushing the system with steam or a cleaner seldom flushes out the stuffing box of the pump.

• Do not circulate shop water through the cooling jacket on a high temperature pump. Condensate or low pressure steam is a better choice. Be sure to install a thermal bushing in the end of the stuffing box to get effective temperature control in the seal area. Make sure you come into the bottom of the jacket and out the top to vent any air that might be trapped in the jacket.

What is the best pump and seal technology?

The "Best Technology" phrase comes up in recent government regulations and every day plant conversations. So what is the best Mechanical Seal and Pump Technology available today? Here is my opinion:

SEAL TECHNOLOGY

Materials

• Identifiable face materials compatible with the fluid to be sealed and any cleaners or solvents put through the lines.

• Materials able to handle the full temperature range of the product you are sealing. • Viton® compatible with water.

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• Hard faces that are not sensitive to temperate change or caustic cleaners. • Unfilled carbon graphite seal faces • No elastomers with shelf life. • No stainless steel springs or bellows.

Design

• The seal should shut with spring and system hydraulic pressure. • Hydraulically balanced designs for low heat generation. • Two way balance in dual seal designs. • Built in pumping ring for cartridge dual seals. • Tandem configuration in dual seal designs. No rotating "back to back" designs. • Stationary configuration for non-cartridge applications. • Self aligning design for stationary cartridge versions. • Springs designed out of the fluid. • The elastomer should move to a clean surface as the faces wear. • No spring loaded elastomers. • Non fretting designs. • Independent of shaft tolerance and finish • Static elastomer located away from the seal face • Cartridge sleeve sealed at wet end. • Vibration damping of the seal face. • Seal should be located close to bearing support. • No elastomer in the seal face. • Faces in compression. • Wide operating range • Low hysteresis. • Equal & opposite clamping of stationary face. • Sealing fluid located at the outside diameter of the seal faces • Leak detection capability • Independent of shaft finish and tolerance • Compensate for thermal expansion and adjustments. • Meet fugitive emission standards. • Simple installation. • Eliminate all elastomers if possible • Short length leaving room for a shaft support bushing. • Finite element analysis of all components. • A method of supporting the shaft in the event of a bearing failure. • Trapped gaskets.

OTHER

• Packaging to survive a one meter drop. • Back up sealing. • Built in seal face vent for vertical applications. • No glued elastomers in split seal configurations.

BEST PUMP TECHNOLOGY

• Low shaft diameter to length ratio (less than 60 L3 /D4) . • Large operating window • C or D frame adapter to simplify driver to pump alignment • Centerline design for thermal expansion. • Oversize stuffing box. • Adequate bearing retention (no snap rings). • Positive bearing sealing. • Oil level indication. • Oil cooling availability. • Low NPSH. • Double volute to prevent shaft deflection. • Suction specific speed number below 8500. • Dynamically balanced rotating assembly. • Impeller specific speed number selected for the application. • Duplex metal impeller. • Impeller investment cast. • Adjust impeller from the wet end to prevent seal face load change.

A new method of troubleshooting centrifugal pumps and mechanical seals:

One of the U. S. based Japanese automobile manufacturers has a unique method of troubleshooting any type of mechanical failure. The system is called the "Five Whys". It is a simple but powerful concept, nothing has been solved until the question "why ?" has been asked at least five times and a sensible answer has been given for each of the "why" questions. As an example:


• The lapped faces opened and solids penetrated between them. (solids can't get in until the faces open)

2. Why did the faces open?

• The set screws holding the rotary unit slipped due to a combination of vibration and system pressure.

3. Set screws are not supposed to slip. Why did the set screws slip?

• The seal was installed on a hardened sleeve.

4. Why was the seal installed on a hardened sleeve?

• This was a packing conversion and a stock sleeve was used.

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5. Why couldn't the mechanic tell the difference between a hardened sleeve and a soft one?

• They were both stored in the same bin.

6 Why were they stored in the same bin?

• Because they had the same part number.

7. Why did they have the same part number?

• They should have had different part numbers. Once that problem is corrected, the failures will stop.

Now you get the idea! Needless to say you may have to go further than just five "whys". Let's try another example:


• The pump was cavitating and the vibration caused the carbon face to crack.

2. Why was the pump cavitating?

• It did not have enough suction head.

3. Why didn't it have enough suction head?

• The level in the tank got too low.

4. Why did the level in the tank get too low?

• I don't know.

You have not finished five "whys" so you better go find out why the level in the tank go too low or the problem is going to repeat its self. In the above example the float got stuck on a corroded rod, giving an incorrect level indication.

One more example should do it:

1. Why did the seal start to leak?

• The elastomer got hard and cracked.

2. Why did the elastomer get hard and crack?

• It got too hot.

3. Why did it get too hot?

• The pump stuffing box ran dry.

4. Why did the stuffing box run dry?

• It was running under a vacuum and it was not supposed to.

5. Why was it running under a vacuum?

• A Goulds pump impeller was adjusted backwards to the back plate and the impeller pump-out rings emptied the stuffing box.

6. Why was it adjusted backwards?

• Most of the pumps in the facility are of the Duriron brand and they normally adjust to the back plate. The mechanic confused the impeller adjustment method. He has since been retrained

This is a powerful trouble shooting technique. I hope you make good use of it.

Operation practices that cause frequent seal and bearing maintenance problems

Wouldn't it be wonderful if the plant operation and maintenance departments could work independently? The fact of the matter is that there are three types of problems we encounter with centrifugal pumps and poor operation is one of them. If you are curious, the other two are design problems and poor maintenance practices.

Seals and bearings account for over eighty five percent (85%) of premature centrifugal pump failure. In the following paragraphs we will be looking at only those operation practices that can, and will cause premature seal and bearing failure. Design and maintenance practices will be discussed in other papers in this series.

When pumps were supplied with jam packing, the soft packing stabilized the shaft to prevent too much deflection. In an effort to save flushing water and to conserve power, many of these same pumps have since been converted to a mechanical seal and the radial stabilization the packing provided has been lost.

The bad operating practices include:

Running the pump dry will cause over-heating and excessive vibration problems that will shorten seal life. Here are some of the common reasons why a pump is run dry:

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• Failing to vent the pump prior to start-up. • Running the tank dry at the end of the operation cycle. • Emptying the tank for steaming or introduction of the next product. • Running on the steam that is being used to flush the tank. • Starting the standby pump without venting it. Venting a hazardous product

can cause a lot of problems with the liquid disposal. Many operators have stopped venting for that reason.

• Tank vents sometimes freeze during cold weather. This will cause a vacuum in the suction tank, and in some cases could collapse the tank.

• Sump fluids are often dirty, corrosive or both. The control rods for the float switch will often "gum up" or corrode and give a false reading to the operator. He may think that there is an adequate level, when in fact, the tank is empty.

Dead heading the pump can cause severe shaft deflection as the pump moves off of its best efficiency point (B.E.P.). This translates to excessive heat that will affect both the seal and the bearings as well as causing the seal faces to open, and the possibility of the impeller contacting the volute when the shaft deflects.

• Starting the centrifugal pump with a shut discharge valve is standard practice with many operation departments. The concern is to save power without realizing the damage that is being done to the mechanical seal, impeller, wear rings and bearings.

• Some pumps are equipped with a recirculation valve that must be opened to lessen the problem, but many times the valve is not opened, or the bypass line is clogged or not of the correct diameter to prevent the excessive head. Another point to remember is that if the bypass line is discharged to the suction side of the pump the increased temperature can cause cavitation.

• After a system has been blocked out the pump is started with one or more valves not opened.

• Discharge valves are shut before the pump has been stopped.

Operating off of the best efficiency point (B.E.P.). Changing the flow rate of the liquid causes shaft deflection that can fail the mechanical seal and over-load the bearings.

• Starting the pump with the discharge valve closed to save power. • The level in the suction tank is changing. Remember that the pump pumps

the difference between the discharge and suction heads. If the suction head varies, the pump moves to a different point on its curve.

• Any upset in the system such as closing, throttling or opening a valve will cause the pump to move to a new point on the curve as the tank fills.

• Pumping to the bottom of a tank will cause the pump to move to a different point on the curve as the tank fills. Some systems were designed for a low capacity positive displacement pump and have since been converted to a

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centrifugal design because of a need for higher capacity. Centrifugal pumps must discharge to the top of the tank to prevent this problem.

• If the discharge piping is restricted because of product build up on the inside walls, the pump will run throttled. This is one of the reasons that it is important to take periodic flow and amperage readings.

• Increasing the flow will often cause cavitation problems.

Seal environmental controls are necessary to insure long mechanical seal life. It is important that operations understand their function and need because many times we find the controls installed, but not functioning.

• Cooling-heating jackets should show a differential temperature between the inlet and outlet lines. If the jacket clogs up, this differential will be lost and seal failure will shortly follow.

• Barrier fluid is circulated between two mechanical seals. There may or may not be a differential temperature depending upon the flow rate. If a convection tank is installed, there should be a temperature differential between the inlet and outlet lines. The line coming out of the top of the seal to the side of the tank should be warmer than the line from the bottom of the tank to the bottom of the seals, otherwise the system is running backwards and may fail completely. The level in the tank is also critical. It should be above the tank inlet line or no convection will occur. Some convection tanks are pressurized with a gas of some type. Many original equipment (O.E.M.) seal designs will fail if this differential pressure is lost.

• Some seal glands (A.P.I. type) are equipped with a quench connection that looks like the seal is leaking water or steam. If there is too much steam pressure on this quench connection, the excessive leakage will get into the bearings causing premature failure. The steam is often used to keep the product warm to prevent it from solidifying, crystallizing, getting too viscous, building a film on the faces etc. Operating people frequently shut off the quench to stop the condensate from leaking.

• Flushing fluids are used for a variety of purposes, but most of the time they are used to get rid of unwanted solids. The flush can be closely controlled with a flow meter or throttling valve. The amount of flush is determined by the seal design. As an example, those designs that have springs in the product require more flush.

• It is important to check that the stuffing box has been vented in vertical pumps. The vent should be coming out of the seal gland and not the stuffing box lantern ring connection.

There are some additional things that all operators should know to insure longer rotating equipment life. As an example :

• Mechanical seals have an 85% or more failure rate that is normally correctable. This is causing unnecessary down time and excessive operating expense. Seals should run until the sacrificial carbon face is

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worn away, but in more that 85% of the cases the seal fails before this happens.

• There are five different causes of cavitation. • You should know where the best efficiency point (B.E.P.) is on a particular

pump, and how far it is safe to operate off the B.E.P. with a mechanical seal installed.

• You should be aware that washing down the pump area with a water hose will cause premature bearing failure when the water penetrates the bearing case.

• Learn about the affect of shaft L3/D4 on pump operation. • Know how the pumped product affects the life of the mechanical seal and

why environmental controls are necessary. • If you are not using cartridge seals, adjusting the open impeller for

efficiency will shorten the seal life. In most cases the seal will open as the impeller is being adjusted to the volute. Durco pumps are the best example of the exception to this rule. The popular Durco pumps adjust to the back plate causing a compression of the seal faces that can create mechanical seal "over heating" problems.

• Cycling pumps for test will often cause a mechanical seal failure unless an environmental control has been installed to prevent the failure.

• Mechanical seals should be positioned after the impeller has been adjusted for thermal growth. This is important on any pump that is operated above 200°F (100°C) or you will experience premature seal failure.

• Some elastomers will be affected by steaming the system. A great deal of caution must be exercised if a flushing fluid such as caustic is going to be circulated through the lines or used to clean a tank. Both the elastomer and some seal faces (reaction bonded silicone carbide is a good example) can be damaged. If the elastomer is attacked, the failure usually occurs within one week of the cleaning procedure.

• The stuffing box must be vented on all vertical centrifugal pumps or otherwise air will be trapped at the seal faces that can cause premature failure of many seal designs.

• Most original equipment seal designs cause shaft damage (fretting) necessitating the use of shaft sleeves that weaken the shaft and restrict pump operation to a narrow range at the B.E.P..

Here are a few common misconceptions that cause friction between maintenance and operation departments

• Shutting the pump discharge valve suddenly, will blow the seal open. • All ceramics cold shock. • High head, low capacity consumes a lot of power. • The pump must come into the shop to change a mechanical seal. • If you use two hard faces or dual mechanical seals in slurry applications,

you will not need flushing water with its corresponding product dilution.

• If you use metal bellows seals for hot oil applications, you will not need the stuffing box cooling jacket operating.

• It is O.K. to use an oversized impeller because throttling back will save power.

A few more thoughts on the subject

• Operators should receive proper schooling on the trouble shooting and maintenance of pumps. In the military and many modern plants, the operator and the maintenance mechanic are often the same person. If the operator knows how the pump works he will have no trouble figuring out the solution to his problem. Too often he is told to keep the flow gage at a certain point, or between two values without understanding what is actually happening with the equipment. If the operator recognizes cavitation he can tell the maintenance department and help them with their trouble shooting.

• As you wander around the plant look out for painters that paint the springs of outside and double mechanical seals. There is a trend to putting two seals in a pump for environmental reasons and the painting of springs is becoming a common problem.

• If someone is adjusting the impeller make sure he is resetting the seal spring tension at the same time.

• If the pump is getting hot or making excessive noises, report it immediately. After the failure, it does no good to tell maintenance that it was making noise for two weeks.

• If you are the floor operator it is common knowledge that taking temperature and pressure readings is very boring, especially on those gages that are located in hot or awkward locations. Avoid the temptation to "radio" these readings. From hot to failure is a very short trip.

• Maintenance's favorite expression is "there is never time to do it right, but there is always time to fix it." Try to keep this in mind when the pressure is on to get the equipment running again.

• Do not let cleaning people direct their "wash down" hoses directly at the pump. Water entering the bearings through the lip or grease seals is a major cause of premature bearing failure. Most water wash downs are used to dilute and wash away seal leakage. Stop the leak and you have eliminated the reason for the hose.

• A great many motor and electrical problems are caused by these same wash down hoses.

• Cooling a bearing outside diameter will cause it to shrink and the bearing will get hotter as the radial load increases. Keep the water hose and all other forms of cooling off of the bearing casing.

The pump is not producing enough head to satisfy the application?

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This is the first paper in a four part series about pump troubleshooting. Let me begin by pointing out that there are a couple of things you must keep in mind when troubleshooting centrifugal pump problems:






THE PROBLEM COULD BE IN THE PUMP ITS SELF

• The impeller diameter is too small. o The impeller is running at too slow a speed o You are running an induction motor. Their speed is different than

synchronous motors. It's always slower. The pump curve was created using a variable frequency motor that ran at a constant speed. Put a tachometer on your motor to see its actual speed.

o Your pulley driven pump is running on the wrong pulley diameter. o A variable frequency motor is running at the wrong speed. o Check the speed of the driver if the pump is driven by something other

than an electric motor. • There is something physically wrong with the motor. Check the bearings etc. • Check the voltage of the electric motor. It may be too low. • The impeller is damaged. The damage could be caused by excessive wear, erosion,

corrosion or some type of physical damage. o Physical damage often occurs during the assembly process when the

impeller is driven on or off the shaft with a wooden block and a mallet. Many impeller designs do not have a nut cast into the impeller hub to ease removal.

o Erosion occurs when solids enter the eye of the impeller. The solids can chip off pieces of the ceramic that are passivating the impeller, causing localized corrosion.

o Damage can occur if the impeller to volute, or back plate clearance is too small and the shaft experiences some type of deflection. The original clearance could have diminished with thermal growth of the shaft. Keep in

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mind that some open impellers adjust to the volute (Goulds) while other designs adjust to the back plate (Duriron).

• In an ANSI and similar design centrifugal pumps, the normal thrust towards the volute has bent the snap ring designed for bearing retention. This can allow the rotating impeller to hit the stationary volute.

• Here are some examples of shaft displacement: o Operating the pump too far off the BEP. o Pulley driven applications. o Pipe strain. o Misalignment between the pump and driver. o The shaft could be bent. o The rotating assembly was probably not dynamically balanced.

• The impeller is clogged. This is a major problem with closed impellers. With the exception of finished product, most of what you will be pumping contains entrained solids. Remember also that some products can solidify, or they can crystallize with a change in fluid temperature or pressure.

• Impeller balance holes have been drilled between the eye and the wear rings of a closed impeller. The reverse flow is interfering with the product entering the impeller eye. A discharge recirculation line should have been used in place of the balance holes to reduce the axial thrust.

• The double volute casting is clogged with solids or solids have built up on the surface of the casting.

• The open impeller to volute clearance is too large. 0.017" (0,5 mm) is typical. This excessive clearance will cause internal recirculation problems. A bad installation, thermal growth, or normal impeller wear could be the cause.

o A large impeller to cutwater clearance can cause a problem called discharge recirculation. Wear is a common symptom of this condition.

• If the impeller is positioned too close to the cutwater you could have cavitation problems that will interfere with the head.

• The impeller specific speed number is too high. Lower specific speed numbered impellers are used to build higher heads.

• An impeller inducer was left off at the time of assembly. Inducers are almost always needed with high specific speed impellers. Leaving off the inducer can cause cavitation problems that will interfere with the head.

• The impeller is loose on the shaft. • The impeller is running backwards • The shaft is running backwards because of a wiring problem. • The pump is running backwards because the discharge check valve is not holding

and system pressure is causing the reverse rotation. This is a common problem with pumps installed in a parallel configuration. Check valves are notoriously unreliable.

• The impeller has been installed backwards. This can happen with closed impellers on double ended pumps

• The second stage of a two stage pump is wired backwards. The pump reverses when the second stage kicks in. You should have heard a loud noise when this happened.

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• The wear ring clearance is too large. o This is a common problem if the shaft L3/D4 number is greater than 60 (2

in the metric system). o You should replace the rings when the original clearance doubles.

Needless to say this can only be determined by inspection. • If you are pumping a product at 200°F (100°C) or more you should use a

centerline design volute to prevent excessive wear ring wear as the volute grows from the base straight up, engaging the wear rings.

• A wear ring is missing. It was probably left off during the installation process. • A high suction tank level is reducing the differential pressure across the pump

increasing its capacity. The pump pumps the difference between the suction and discharge heads.

• A bubble is trapped in the eye of the impeller. The eye is the lowest pressure area. When this bubble forms it shuts off all liquid coming into the pump suction. This could cause the pump to lose its prime.

• You cannot vent a running pump because centrifugal force will throw the liquid out the vent leaving the air trapped inside.

• Air is coming directly into the pump. This happens with a negative pressure at the suction side. Negative suction happens when the pump is lifting liquid, pumping from a condenser hot well etc.

o Air is coming into the stuffing box through the pump packing. o Air is coming into the stuffing box through an unbalanced mechanical seal.

As the carbon face wears the spring load holding the faces together diminishes.

o If you are using mechanical seals in vacuum service, they should be of the O-ring design. Unlike other designs, O-rings are the only shape that seals both pressure and vacuum.

o The pump was not primed prior to start up. With the exception of the self priming version, centrifugal pumps must be full of liquid at start up.

o Air can enter the stuffing box if the gasket between the two halves of a double ended pump is defective or does not extend to the stuffing box face. Any small gaps between the face of the stuffing box and the split at the side of the stuffing box will allow either air in, or product out.

o Air is coming into the suction side of the pump through a pin hole in the casing.

o Air is entering the stuffing box between the sleeve and the shaft. This happens if you convert a double ended pump from packing to a mechanical seal and fail to install a gasket or o-ring between the impeller hub and the sleeve.

• The open impeller was adjusted backwards and now the close fitting "pump out vanes" are creating a vacuum in the stuffing box.

• You need a volute casing instead of a concentric casing. Volute casings are much better for producing head.

• You have the wrong size pump. It cannot meet the system curve requirements: • The pump was not selected to meet the system curve requirements because no

system curve was given to the pump supplier.

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• At replacement time the same size pump was purchased because no one had calculated losses in the system.

• The pump was sized from a piping diagram that was thirty five years old. There have been numerous piping changes and additions since the original layout. In many instances additional pumps have been installed and this pump is running in parallel with them, but nobody knows it.

THE PROBLEM IS ON THE SUCTION SIDE OF THE PUMP. THE PUMP COULD BE CAVITATING.

• Air is entering the suction piping at some point. o Air is being pumped into the suction piping to reduce cavitation problems o Fluid returning to the sump is being aerated by too far a free fall. The

return line should terminate below the liquid level. o The fluid is vortexing at the pump inlet because the sump level is too low

and the pump capacity is too high. o Air is coming into the system through valves above the water line or

gaskets in the piping flanges. o The liquid source is being pumped dry. If this is a problem in your

application you might want to consider a self priming pump in the future. • The vapor pressure of the fluid is too close to atmospheric pressure. When it rains

the drop in atmospheric pressure causes the inlet fluid to vaporize. • There is a problem with the piping layout. It is reducing the head on the suction

side of the pump. o There is too much piping between the pump suction and the source tank.

You may need a booster pump or an inducer. The higher the pump speed the bigger the problem.

o There is an elbow too close to the pump suction. There should be at least ten diameters of pipe between the elbow and the pump suction. Suction piping should never run parallel with the pump shaft in a double ended pump installation. This can cause unnecessary shaft thrusting.

o A piece of pipe of reduced diameter has been installed in the suction piping.

o Piping was added on the inlet side of the pump to by-pass a piece of equipment that was installed on the floor.

o A piping to pump reducer has been installed upside down causing an air pocket. Concentric reducers can cause the same problem..

o Multiple pump inlets are too close together. • The pump inlet is too close to the tank floor. • The suction lift is too high. • A gasket with too small an inside diameter has been installed in the suction piping

restricting the liquid flow. • A gasket in the suction piping is not centered and is protruding into the product

stream. • A globe valve has been substituted for a gate valve in the suction piping. The loss

of head in a globe valve is many times that of a gate valve.

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• Two pumps are connected in series. The first pump is not sending enough capacity to the second pump.

• The piping inlet is clogged. • A filter or strainer is clogged or covered with something. • Intermittent plugging of the suction inlet.

o Loose rags can do this. o If the suction is from a pond, river, or the sea, grass can be pulled into the

suction inlet. • A foot valve is stuck. • A check valve is stuck partially closed • The foot valve is too small. • A small clam or marine animal cleared the suction screen, but has now grown

large on the pump side of the screen. • The suction piping diameter has been reduced.

o The suction piping collapsed when a heavy object either hit or ran over the piping.

o Solids have built up on the piping walls. Hard water is a good example of this problem

o A liner has broken away from the piping wall and has collapsed in the piping. Look for corrosion in the piping caused by a hole in the liner.

o A foreign object is stuck in the piping It was left there when the piping was repaired.

o The suction is being throttled to prevent the heating of the process fluid. This is a common operating procedure with fuel pumps where discharge throttling could cause a fire or explosion.

• The pump inlet temperature is too high. o The tank is being heated to deaerate the fluid, but it is heating the fluid up

too much. Look for this problem in boiler feed pump applications. o The sun is heating the inlet piping. The piping should be insulated to

prevent this problem. o The operating temperature of the pumped fluid has been increased to

accommodate the process requirements. o A discharge recirculation line is heating the incoming fluid. You should

direct this line to a reservoir rather than the pump suction. o Steam or some other hot cleaner is being circulated through the lines.

• The problem is in the tank connected to the suction of the pump. o The pump capacity is too high for the tank volume. o The tank float is stuck, showing a higher tank level that does not exist. o The tank vent is partially shut or frozen, lowering the suction pressure. o There is not enough NPSH available for the fluid you are pumping. Maybe

you can use an inducer or booster pump to increase the suction pressure. o A high suction tank level is reducing the differential pressure across the

pump, increasing its capacity and lowering the head.


• Two pumps are in connected in series. The first pump does not have enough capacity for the second pump. They should be running at the same speed with the same width impeller.

• The pump discharge is connected to the bottom of the tank. The head is low until the level in the tank increases.

• Units in the discharge piping should not normally be shut off, they should be by-passed to prevent too much of a change in the pump's capacity.

• If too many units are being by-passed in the discharge system the head will decrease as the capacity increases. This can happen if an extra storage tank farm is being by-passed because the storage capacity is no longer needed.

• A bypass line has been installed in the pump discharge increasing the capacity and lowering the head.

• Piping or fittings have been removed from the discharge side of the pump reducing piping resistance.

• Connections have been installed in the discharge piping that have increased the demand that increases capacity.

• The pump is acting as an accumulator, coming on when the tank level drops. The head will be low until the accumulator is recharged.

• Consider the possibility of a siphon affect in the discharge piping. This will occur if the pump discharge piping is entering into the top of a tank and discharging at a lower level The pump must build enough head initially to take advantage of the siphoning action.

• A discharge valve (manual or automatic) is opened too much.

The pump is not producing enough capacity to satisfy the application?

Let me begin by pointing out that there are a couple of things you must keep in mind when troubleshooting centrifugal pump problems:





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THE PROBLEM IS IN THE PUMP ITS SELF:

• The impeller diameter is too small • The impeller width is too narrow • The impeller speed is too slow. Check the voltage and frequency • The impeller is damaged. • The impeller is clogged. • The open impeller clearance is too large. • The impeller to cutwater clearance is too large. • The impeller specific speed number is too low. • The impeller has been installed backwards • The shaft is running backwards. • The wear ring clearance is too large. • A wear ring is missing. • The second stage of a two stage pump is wired backwards. • A bubble is trapped in the eye of the impeller. • A low suction tank level is increasing the differential pressure across the pump

decreasing its capacity. • Air is coming into the pump suction through the packing. • Air is coming into the pump suction through an unbalanced mechanical seal. • The pump was not primed prior to star up. • You may need a concentric casing rather than the volute design. • You are using a variable speed motor trying to produce a flat curve. Remember

that both the head and capacity change with speed. • The pump is the wrong size. Someone gave the pump distributor a wrong system

curve

THE PROBLEM IS ON THE SUCTION SIDE OF THE PUMP

• There is too much piping between the pump suction and the source tank. • There is an elbow too close to the pump suction. • A filter or strainer is clogged. • Intermittent plugging of the suction inlet. Loose rags can do this. • A foot valve is stuck • The tank float is stuck. Showing a higher tank level that does not exist. • The tank vent is partially shut or frozen. • A globe valve has been substituted for a gate valve. • A check valve is stuck partially closed • Solids have built up on the piping walls. • A liner has broken away from the piping wall and has collapsed in the piping. • The piping was collapsed by a heavy object that hit the outside of the piping. • A foreign object is stuck in the piping It was left there when the piping was

repaired.

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• A small clam cleared the suction screen, but has now grown large on the pump side of the screen.

• The sun is heating the inlet piping. It should be insulated to prevent this problem. • Piping was added on the inlet side of the pump to compensate for a piece of

equipment that was installed in the shop. • A reducer has been installed upside down. • A discharge recirculation line is heating the incoming fluid. • The pump capacity is too high for the tank volume. • Multiple pump inlets are too close together. • The suction lift is too high. • There is not enough NPSH available for the fluid you are pumping. Maybe you

can use an inducer to increase the suction pressure. • Air is coming into the system through valves above the water line or gaskets in

the piping. • Air is being pumped into the suction piping to reduce cavitation problems • Fluid returning to the sump is being aerated by too far a free fall. • The fluid is vortexing at the pump inlet because the sump level is too low. • The tank is being heated to deaerate the fluid, but it is heating the fluid up too

much. • Two pumps are connected in series. The first pump is not sending enough

capacity to the second pump. • The operating temperature of the pumped fluid has increased. • The vapor pressure of the fluid is too close to atmospheric pressure. When it rains

the drop in atmospheric pressure causes the inlet fluid to vaporize. • The suction is being throttled to prevent the heating of the process fluid.


• Extra piping has been added to the system to accommodate extra storage capacity. • A bypass line has been installed in the pump discharge. • Piping or fittings have been added to the discharge side of the pump. • An orifice has been installed in the discharge piping to reduce the capacity or

produce a false head. • A gate valve has been substituted for a globe valve in the discharge piping. • A check valve is stuck partially closed. • An orifice has been installed into the piping to restrict flow. • The piping was collapsed by a heavy object that hit the outside of the piping. • The discharge valve is throttled too much. • There is a restriction in the discharge piping. • Extra pumps have been installed into the existing piping They are connected in

parallel, but are not producing the same head. • Two pumps are in parallel. The larger one is shutting the check valve of the

smaller pump. • Two pumps are in connected in series. The first pump does not have enough

capacity for the second pump. They should be running at the same speed with the same width impeller

• The pump discharge is connected to the bottom of the tank. The head is increasing and the capacity is decreasing as the tank fills.

• The pump is acting as an accumulator&emdash;coming on when the tank level drops. The head is too high when the tank fills.

Pump selection practices that cause high seal and bearing maintenance problems:

Purchasing well designed hardware does not bring automatic trouble free performance with it. The very best equipment will cause problems if it was not designed for your particular application. Here are a few of the more common selection problems we find with centrifugal pumps:

• Buying the same size pump as the one that came out of the application. That's O.K. If the old pump was the correct size, but the odds are that it was too big because of the safety factors that were added at the time of purchase. This will cause the pump to run off of its best efficiency point (B.E.P.) and you will spend a lot of production money for the additional power that is needed to run against a throttled discharge valve or orifice installed in the discharge piping.

• Buying to a standard, or making a decision based on efficiency, and believing that these two some how relate to quality. Standards were written for packed pumps. When a mechanical seal is being used the shaft L3/D4 number is almost always too large. Efficiency is always gained at the expense of maintenance. Efficiency means tight tolerances and smooth passages that will eliminate reliable double volute designs and keep the maintenance department busy adjusting tight tolerances to maintain the efficiency you paid for.

• Series and parallel installation problems. We often find pumps installed in parallel, but no one knows it because the second pump was installed at a much later date and no one has bothered to trace the piping. Pumps in parallel require that they have the same diameter impeller and that they run at the same speed, or the larger pump will throttle the smaller one causing it to run off the best efficiency point, deflecting the shaft. The capacity should be looked at if the higher capacity pump might exceed the N.P.S.H. available.

• When pumps are installed in series the impellers must be the same width and they must run at the same speed or the higher capacity pump will either cavitate because the smaller capacity pump can not feed liquid at the proper capacity, or it will run throttled if it is feeding the smaller pump. In either case the larger of the two pumps will be adversely affected.

• Purchasing a larger pump because it will be needed in the future. Will raise the operating cost to unacceptable levels (Power = head x capacity) as the pump is run against a throttled discharge valve. This inefficient use of power will translate to a higher heat environment for the seal along with all of the problems associated with shaft deflection.

• Using a variable speed motor to compensate for a pump curve that is not flat enough. Many boiler feed pumps require a flat curve so that the pump can put out

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varying capacities at a constant boiler pressure (head). We see this same need if we are pumping a varying amount of liquid to a very high constant height.

• Varying the speed of a pump is similar to changing the diameter of the impeller. If you look at a typical pump curve you will observe that the best efficiency point (B.E.P.) comes down with impeller size to form an angle with the base line (capacity line) of the graph. This means that if you vary the speed of the impeller, the pump always runs off the B.E.P. except in the case where the system curve intersects the pump curve, or in the case of an exponential system curve such as we find in a typical hot or cold water circulating system.

• Double ended pumps installed in a vertical position to save floor space. Makes seal replacement a nightmare unless you are using split or cartridge designs.

• Specifying a desired capacity without knowing the true system head. You can't guess with this one. Some one has to make the calculations and "walk the system". The present pump is not a reliable guide because we seldom know where it is pumping on its' curve. Chart recorders installed on both the suction and discharge side of the pump will give a more accurate reading of the present head if they are left on long enough to record the differences in flow. The trouble with this method is that it will also record a false head caused by a throttled valve, an orifice, or any other restriction that might be present in the piping.

• Requesting too low a required N.P.S.H. will cause you to end up with a different kind of cavitation problem. See another paper in this series for information about "Internal recirculation".

• Failure to request a "center line design" when pumping temperature exceeds 200°F (100°C) it will cause pipe strain that will translate to wear ring damage and excessive mechanical seal movement.

• The use of "inline" pumps to save floor space. Many of these designs are "close coupled" with the motor bearings carrying the radial and thrust loads. Because of typical L3/D4 numbers being very high, the wear rings act as "steady bearings" after the pump is converted to a mechanical seal. The pump should have been designed with a separate bearing case and a "C" or "D" frame adapter installed to connect a motor to the bearing case.

• Thrust bearings being retained by a simple snap ring. Beyond 65% of its rated efficiency most centrifugal pumps thrust towards the pump volute. The thin snap ring has to absorb all of this axial thrust and most of them can not do it very well .

• The mechanical seal has been installed in a packing stuffing box that is too narrow to allow free seal movement. If a mechanical seal was specified, the pump back plate should have been manufactured with a large diameter seal chamber. In most cases the stuffing box recirculation line should be installed from the bottom of this large seal chamber to the suction side of the pump or a low pressure point in the system. There are some exceptions to this, however:

o If you are pumping at or close to vapor point. o If the entrained solids have a low specific gravity. o If you are using a Duriron pump that adjusts to the back plate. o If you are using a double suction pump where the stuffing boxes are at

suction pressure. • High temperature applications have several special needs:

o A jacketed stuffing box that isolates the pumpage from the stuffing box contents by a carbon bushing to retard heat transfer.

o A centerline design to compensate for thermal expansion. o A cartridge seal design that allows open impeller adjustment after the

pump has come up to operating temperature. o A stainless steel shaft to retard heat transfer to the bearings. o A method of cooling the bearing oil, but never the bearings. o A coupling that will compensate for axial expansion.

Pipe Data Carbon and Alloy Steel - Stainless Steel

Pipe Size Inches

Outside

Diam.

Inches

Identification

Wall Thickness (t)Inches

Inside Diameter (d)Inches

Area of

Metal

Square

Inches

Transverse Internal

AreaMoment of

Inertia (l)

Inches

Weight

PipePoun

ds per foot

Weight

Water

Pounds per foot

External

Surface

Sq. Ft. per foot of

pipe

Section

Modulus

Steel Stainle

ss Steel

Sched. No.

(a)Squa

re Inches

(A)Square Feet

Iron

Pipe

Size

Sched.

No.

1/8 0.405

STD

XS

40 80

10S 40S 80S

0.490.680.95

.307

.269

.215

.0548

.0720

.0925

.0740

.0568

.0364

.00051

.00040

.00025

.00088

.00106

.00122

.19

.24

.31

.032

.025

.016

.106

.106

.106

.00437

.00523

.00602

¼ 0.540

STD

XS

40 80

10S 40S 80S

.065

.088

.119

.410

.364

.302

.0970

.1250

.1574

.1320

.1041

.0716

.00091

.00072

.00050

.00279

.00331

.00377

.33

.42

.54

.057

.045

.031

.141

.141

.141

.01032

.01227

.01395

3/8 0.675

STD

XS

40 80

10S 40S 80S

.065

.091

.126

.545

.493

.423

.1246

.1670

.2173

.2333

.1910

.1405

.00162

.00133

.00098

.00586

.00729

.00862

.42

.57

.74

.101

.083

.061

.178

.178

.178

.01736

.02160

.02554

½ 0.840

ST

40 80

5S 10S 40S

.065

.083

.109

.710

.674

.622

.1583

.197

.3959

.356

.00275

.0024

.01197

.0143

.54

.67

.85

.172

.155

.132

.220

.220 220

.02849

.03407

.04069

155

156

D XS

XXS

160 80S

.147

.187

.294

.546

.466

.252

4 .250

3 .320

0 .383

6 .504

3

8 .304

0 .234

0 .170

6 .050

8 .0021

1 .0016

3 .0011

8 .0003

5

1 .0170

9 .0200

8 .0221

2 .0242

4

1.091.311.71

.102

.074

.022

.220 220 .220

.04780

.05267

.05772

¾ 1.050

STD

XS

XXS

40 80 160

5S 10S 40S 80S

.065

.083

.113

.154

.219

.308

.920

.884

.824

.742

.612

.434

.2011

.2521

.3326

.4335

.5698

.7180

.6648

.6138

.5330

.4330

.2961

.148

.00462

.00426

.00371

.00300

.00206

.00103

.02450

.02969

.03704

.04479

.05269

.05792

.69

.861.131.471.942.44

.288

.266

.231

.188

.128

.064

.275

.275

.275

.275

.275

.275

.04667

.05655

.07055

.08531

.10036

.11032

1 1.315

STD

XS

XXS

40 80 160

5S 10S 40S 80S

.065

.109

.133

.179

.250

.358

1.1851.0971.049.957.815.599

.2553

.4130

.4939

.6388

.8365

1.0760

1.1029

.9452

.8640

.7190

.5217

.282

.00766

.00656

.00600

.00499

.00362

.00196

.04999

.07569

.08734

.1056

.1251

.1405

.871.401.682.172.843.66

.478

.409

.375

.312

.230

.122

.344

.344

.344

.344

.344

.344

.07603

.11512.1328.1606.1903.2136

1 ¼ 1.660

STD

XS

XXS

40 80 160

5S 10S 40S 80S

.065

.109

.140

.191

.250

.382

1.5301.4421.3801.2781.160.896

.3257

.4717

.6685

.8815

1.1070

1.534

1.839

1.633

1.495

1.283

1.057

.630

.01277

.01134

.01040

.00891

.00734

.00438

.1038

.1605

.1947

.2418

.2839

.3411

1.111.812.273.003.765.21

.797

.708

.649

.555

.458

.273

.435

.435

.435

.435

.435

.435

.1250

.1934

.2346

.2913

.3421

.4110

1 ½ 1.900

STD

XS

XXS

40 80 160

5S 10S 40S 80S

.065

.109

.145

.200

.281

.400

1.7701.6821.6101.5001.3381.100

.3747

.6133

.7995

1.068

1.42

2.461

2.222

2.036

1.767

1.40

.01709

.01543

.01414

.01225

.0097

.1579

.2468

.3099

.3912

.4824

.5678

1.282.092.723.634.866.41

1.066 .963 .882 .765 .608 .42

.497

.497

.497

.497

.497

.497

.1662

.2598

.3262

.4118

.5078

.5977

157

9 1.88

5

6 .950

6 .0066

0

2 2.375

STD

XS

XXS

40 80 160

5S 10S 40S 80S

.065

.109

.154

.218

.344

.436

2.2452.1572.0671.9391.6871.503

.4717

.7760

1.075

1.477

2.190

2.656

3.958

3.654

3.355

2.953

2.241

1.774

.02749

.02538

.02330

.02050

.01556

.01232

.3149

.4992

.6657

.86791.1621.311

1.612.643.655.027.469.03

1.72 1.58 1.45 1.28 .97 .77

.622

.622

.622

.622

.622

.622

.2652

.4204

.5606

.7309.979

1.104

2 ½ 2.875

STD

XS

XXS

40 80 160

5S 10S 40S 80S

.083

.120

.203

.276

.375

.552

2.7092.6352.4692.3232.1251.771

.7280

1.039

1.704

2.254

2.945

4.028

5.764

5.453

4.788

4.238

3.546

2.464

.04002

.03787

.03322

.02942

.02463

.01710

.7100

.98731.5301.9242.3532.871

2.483.535.797.66

10.01 13.69

2.50 2.36 2.07 1.87 1.54 1.07

.753

.753

.753

.753

.753

.753

.4939

.68681.0641.3391.6381.997

3 3.500

STD

XS

XXS

40 80 160

5S 10S 40S 80S

.083

.120

.216

.300

.438

.600

3.3343.2603.0682.9002.6242.300

.8910

1.274

2.228

3.016

4.205

5.466

8.730

8.347

7.393

6.605

5.408

4.155

.06063

.05796

.05130

.04587

.03755

.02885

1.3011.8223.0173.8945.0325.993

3.034.337.58

10.25 14.32 18.58

3.78 3.62 3.20 2.6

2.35 1.80

.916

.916

.916

.916

.916

.916

.74351.0411.7242.2252.8763.424

3 ½ 4.000

STD

XS

40 80

5S 10S 40S 80S

.083

.120

.226

.318

3.8343.7603.5483.364

1.021

1.463

2.680

3.678

11.545

11.104

9.886

8.888

.08017

.07711

.06870

.06170

1.9602.7554.7886.280

3.484.979.11

12.50

5.00 4.81 4.29 3.84

1.047 1.047 1.047 1.047

.97991.3782.3943.140

4 4.500

STD

NS

40 80 120 160

5S 10S 40S 80S

.083

.120

.237

.337

.438

.531

4.3344.2604.0263.8263.6243.438

1.152

1.651

3.174

14.75

14.25

12.73

.10245

.09898

.08840

2.8103.9637.2339.61011.6513.27

3.925.61

10.79 14.98 19.0

22.51

6.39 6.18 5.50 4.98 4.47 4.02

1.178 1.178 1.178 1.178 1.178 1.178

1.2491.7613.2144.2715.1785.898

158

XXS

.674 3.152 4.407

5.595

6.621

8.101

11.50

10.31

9.287.80

.07986

.0716

.0645

.0542

15.28 27.54 3.38 1.178 6.791

5 5.563

STD

XS

XXS

40 80 120 160

5S 10S 40S 80S

.109

.134

.258

.375

.500

.625

.750

5.3455.2955.0474.8134.5634.3134.063

1.868

2.285

4.300

6.112

7.953

9.696

11.340

22.44

22.02

20.01

18.19

16.35

14.61

12.97

.1558

.1529

.1390

.1263

.1136

.1015

.0901

6.9478.42515.1620.6725.7330.0333.63

6.367.77

14.62 20.78 27.04 32.96 38.55

9.72 9.54 8.67 7.88 7.09 6.33 5.61

1.456 1.456 1.456 1.456 1.456 1.456 1.456

2.4983.0295.4517.4319.25010.79612.090

6 6.625

STD

XS

XXS

40 80 120 160

5S 10S 40S 80S

.109

.134

.280

.432

.562

.719

.864

6.4076.3576.0655.7615.5015.1874.897

2.231

2.733

5.581

8.405

10.70

13.32

15.64

32.24

31.74

28.89

26.07

23.77

21.15

18.84

.2239

.2204

.2006

.1810

.1650

.1469

.1308

11.8514.4028.1440.4949.6158.9766.33

7.609.29

18.97 28.57 36.39 45.35 53.16

13.97 13.75 12.51 11.29 10.30 9.16 8.16

1.734 1.734 1.734 1.734 1.734 1.734 1.734

3.5764.3468.49612.2214.9817.8120.02

8 8.625

STD

XS

XXS

20 30 40 60 80 100 120 140

160

5S 10S

40S

80S

.109

.148

.250

.277

.322

.406

.500

.594

.719

.812

.875

.906

8.4078.3298.1258.0717.9817.8137.6257.4377.1877.0016.8756.813

2.916

3.941

6.577.268.4010.4

8 12.7

6 14.9

6 17.8

4 19.9

3 21.3

55.51

54.48

51.85

51.16

50.03

47.94

45.66

43.46

40.59

.3855

.3784

.3601

.3553

.3474

.3329

.3171

.3018

.2819

.2673

.2578

.2532

26.4435.4157.7263.3572.4988.73105.7121.3140.5153.7162.0165.9

9.9313.40 22.36 24.70 28.55 35.64 43.39 50.95 60.71 67.76 72.42 74.69

24.06 23.61 22.47 22.17 21.70 20.77 19.78 18.83 17.59 16.68 16.10 15.80

2.258 2.258 2.258 2.258 2.258 2.258 2.258 2.258 2.258 2.258 2.258 2.258

6.1318.21213.3914.6916.8120.5824.5128.1432.5835.6537.5638.48

159

0 21.9

7

38.50

37.12

36.46

10 10.750

STD

XS

XXS

20 30 40 60 80 100 120 140 160

5S 10S

40S 80S

.134

.165

.250

.307

.365

.500

.594

.719

.8441.0001.125

10.48210.42010.25010.13610.0209.7509.5629.3129.0628.7508.500

4.365.498.2410.0

7 11.9

0 16.1

0 18.9

2 22.6

3 26.2

4 30.6

3 34.0

2

86.29

85.28

82.52

80.69

78.86

74.66

71.84

68.13

64.53

60.13

56.75

.5992

.5922

.5731

.5603

.5475

.5185

.4989

.4732

.4481

.4176

.3941

63.076.9113.7137.4160.7212.0244.8286.1324.2367.8399.3

15.19 18.65 28.04 34.24 40.48 54.74 64.43 77.03 89.29 104.1

3 115.6

4

37.39 36.95 35.76 34.96 34.20 32.35 31.13 29.53 27.96 26.06 24.59

2.814 2.814 2.814 2.814 2.814 2.814 2.814 2.814 2.814 2.814 2.814

11.7114.3021.1525.5729.9039.4345.5453.2260.3268.4374.29

12 12.75

STD

XS

XXS

20 30

40

60 80 100 120 140 160

5S 10S

40S

80S

.156

.180

.250

.330

.375

.406

.500

.562

.688

.8441.0001.1251.312

12.43812.39012.25012.09012.00011.93811.75011.62611.37411.06210.75010.50010.126

6.177.119.8212.8

7 14.5

8 15.7

7 19.2

4 21.5

2 26.0

3 31.5

3 36.9

1 41.0

8 47.1

4

121.50

120.57

117.86

114.80

113.10

111.93

108.43

106.16

101.64

96.14

90.76

86.59

80.53

.8438

.8373

.8185

.7972

.7854

.7773

.7528

.7372

.7058

.6677

.6303

.6013

.5592

122.4140.4191.8248.4279.3300.3361.5400.4475.1561.6641.6700.5781.1

20.98 24.17 33.38 43.77 49.56 53.52 65.42 73.15 88.63 107.3

2 125.4

9 139.6

7 160.2

7

52.65 52.25 51.07 49.74 49.00 48.50 46.92 46.00 44.04 41.66 39.33 37.52 34.89

3.338 3.338 3.338 3.338 3.338 3.338 3.338 3.338 3.338 3.338 3.338 3.338 3.338

19.222.030.239.043.847.156.762.874.688.1

100.7109.9122.6

160

14 14.00

STD

XS

10 20 30 40

60 80 100 120 140 160

5S 10S

.156

.188

.250

.312

.375

.438

.500

.594

.750

.9381.0941.2501.406

13.68813.62413.50013.37613.25013.12413.00012.81212.50012.12411.81211.50011.188

6.788.1610.8

0 13.4

2 16.0

5 18.6

6 21.2

1 24.9

8 31.2

2 38.4

5 44.3

2 50.0

7 55.6

3

147.15

145.78

143.14

140.52

137.88

135.28

132.73

128.96

122.72

115.49

109.62

103.87

98.31

1.0219

1.0124

.9940

.9758

.9575

.9394

.9217

.8956

.8522

.8020

.7612

.7213

.6827

162.6194.6255.3314.4372.8429.1483.8562.3678.3824.4929.61027.

0 1117.

0

23.07 27.73 36.71 45.61 54.57 63.44 72.09 85.05 106.1

3 130.8

5 150.7

9 170.2

8 189.1

1

63.77 63.17 62.03 60.89 59.75 58.64 57.46 55.86 53.18 50.04 47.45 45.01 42.60

3.665 3.665 3.665 3.665 3.665 3.665 3.665 3.665 3.665 3.665 3.665 3.665 3.665

23.227.836.645.053.261.369.180.398.2

117.8132.8146.8159.6

16 16.00

STD

XS

10 20 30 40 60 80 100 120 140 160

5S 10S

.165

.188

.250

.312

.375

.500

.656

.8441.0311.2191.4381.594

15.67015.62415.50015.37615.25015.00014.68814.31213.93813.56213.12412.812

8.219.3412.3

7 15.3

8 18.4

1 24.3

5 31.6

2 40.1

4 48.4

8 56.5

6 65.7

8 72.1

0

192.85

191.72

188.69

185.69

182.65

176.72

169.44

160.92

152.58

144.50

135.28

128.96

1.3393

1.3314

1.3103

1.2895

1.2684

1.2272

1.1766

1.1751.059

6 1.003

5 .9394.8956

257.3291.9383.7473.2562.1731.9932.41155.

8 1364.

5 1555.

8 1760.

3 1893.

5

27.90 31.75 42.05 52.27 62.58 82.77 107.5

0 136.6

1 164.8

2 192.4

3 223.6

4 245.2

5

83.57 83.08 81.74 80.50 79.12 76.58 73.42 69.73 66.12 62.62 58.64 55.83

4.189 4.189 4.189 4.189 4.189 4.189 4.189 4.189 4.189 4.189 4.189 4.189

32.236.548.059.270.391.5

116.6144.5170.5194.5220.0236.7

18 18.00

STD

10 20

5S 10S

.165

.188

.250

.312

.375

17.67017.62417.50017.37617.250

9.2510.5

2 13.9

4

245.22

243.95

240.

1.7029

1.6941

1.670

367.6417.3549.1678.2806.7

31.43 35.76 47.39 58.94 70.59

106.26

105.71

104.2

4.712 4.712 4.712 4.712 4.712

40.846.461.175.589.6

161

XS

30

40 60 80 100 120 140 160

.438

.500

.562

.750

.9381.1561.3751.5621.781

17.12417.00016.87616.50016.12415.68815.25014.87614.438

17.34

20.76

24.17

27.49

30.79

40.64

50.23

61.17

71.81

80.66

90.75

53237.13

233.71

230.30

226.98

223.68

213.83

204.24

193.30

182.66

173.80

163.72

3 1.646

7 1.623

0 1.599

0 1.576

3 1.553

3 1.484

9 1.418

3 1.342

3 1.268

4 1.207

0 1.136

9

930.31053.

2 1171.

5 1514.

7 1833.

0 2180.

0 2498.

1 2749.

0 3020.

0

82.15 93.45 104.6

7 138.1

7 170.9

2 207.9

6 244.1

4 274.2

2 308.5

0

1 102.7

7 101.1

8 99.84 98.27 96.93 92.57 88.50 83.76 79.07 75.32 70.88

4.712 4.712 4.712 4.712 4.712 4.712 4.712 4.712 4.712

103.4117.0130.1168.3203.8242.3277.6305.5335.6

20 20.00

STD

XS

10 20 30 40 60 80 100 120 140 160

5S 10S

.188

.218

.250

.375

.500

.594

.8121.0311.2811.5001.7501.969

19.62419.56419.50019.25019.00018.81218.37617.93817.43817.00016.50016.062

11.70

13.55

15.51

23.12

30.63

36.15

48.95

61.44

75.33

87.18

100.33

111.49

302.46

300.61

298.65

290.04

283.53

278.00

265.21

252.72

238.83

226.98

213.82

202.67

2.1004

2.0876

2.0740

2.0142

1.9690

1.9305

1.8417

1.7550

1.6585

1.5762

1.4849

1.4074

574.2662.8765.41113.

0 1457.

0 1703.

0 2257.

0 2772.

0 3315.

2 3754.

0 4216.

0 4585.

5

39.78 46.06 52.73 78.60 104.1

3 123.1

1 166.4

0 208.8

7 256.1

0 296.3

7 341.0

9 379.1

7

131.06

130.27

129.42

125.67

122.87

120.46

114.92

109.51

103.39

98.35 92.66 87.74

5.236 5.236 5.236 5.236 5.236 5.236 5.236 5.236 5.236 5.236 5.236 5.236

57.466.375.6

111.3145.7170.4225.7277.1331.5375.5421.7458.5

22 22.00

STD

XS

10 20 30 60 80 100

5S 10S

.188

.218

.250

.375

.500

.8751.1251.375

21.62421.56421.50021.25021.00020.25019.7519.25

12.88

14.92

17.08

25.48

367.25

365.21

363.05

354.66

2.5503

2.5362

2.5212

2.4629

766.2884.81010.

3 1489.

7 1952.

5

43.80 50.71 58.07 86.61 114.8

1 197.4

1

159.14

158.26

157.32

153.68

5.760 5.760 5.760 5.760 5.760 5.760 5.760 5.760

69.780.491.8

135.4117.5295.0366.4432.6

162

120 140 160

1.6251.8752.125

18.7518.2517.75

33.77

58.07

73.78

89.09

104.02

118.55

132.68

346.36

322.06

306.35

291.04

276.12

261.59

247.45

2.4053

2.2365

2.1275

2.0211

1.9175

1.8166

1.71840

3244.9

4030.4

4758.5

5432.0

6053.7

6626.4

250.81

302.88

353.61

403.00

451.06

150.09

139.56

132.76

126.12

119.65

113.36

107.23

5.760 5.760 5.760

493.8550.3602.4

24 24.00

STD

XS

10 20

30 40 60 80 100 120 140 160

5S 10S

.218

.250

.375

.500

.562

.688

.9691.2191.5311.8122.0622.344

23.56423.50023.25023.00022.87622.62422.06221.56220.93820.37619.87619.312

16.29

18.65

27.83

36.91

41.39

50.31

70.04

87.17

108.07

126.31

142.11

159.41

436.10

433.74

424.56

415.48

411.00

402.07

382.35

365.22

344.32

326.08

310.28

292.98

3.0285

3.0121

2.9483

2.8853

2.8542

2.7921

2.6552

2.5362

2.3911

2.2645

2.1547

2.0346

1151.6

1315.4

1942.0

2549.5

2843.0

3421.3

4652.8

5672.0

6849.9

7825.0

8625.0

9455.9

55.37 63.41 94.62 125.4

9 140.6

8 171.2

9 238.3

5 296.5

8 367.3

9 429.3

9 483.1

2 542.1

3

188.98

187.95

183.95

179.87

178.09

174.23

165.52

158.26

149.06

141.17

134.45

126.84

6.283 6.283 6.283 6.283 6.283 6.283 6.283 6.283 6.283 6.283 6.283 6.283

96.0109.6161.9212.5237.0285.1387.7472.8570.8652.1718.9787.9

26 26.00

STD

XS

10

20

.312.375.500

25.37625.25025.000

25.18

30.19

40.06

505.75

500.74

490.87

3.5122

3.4774

3.4088

2077.2

2478.4

3257.0

85.60 102.6

3 136.1

7

219.16

216.99

212.71

6.806 6.806 6.806

159.8190.6250.5

28 28.00

STD

XS

10

20 30

.312.375.500.625

27.37627.25027.00026.750

27.14

32.54

43.20

53.75

588.61

583.21

572.56

562.00

4.0876

4.0501

3.9761

3.9028

2601.0

3105.1

4084.8

5037.7

92.26 110.6

4 146.8

5 182.7

3

255.07

252.73

248.11

243.53

7.330 7.330 7.330 7.330

185.8221.8291.8359.8

30 30.00

10

5S 10S

.250

.31229.50029.376

23.37

683.49

4.7465

2585.2

79.43 98.93

296.18

7.854 7.854

172.3213.8

STD

XS

20 30

.375

.500

.625

29.25029.00028.750

29.10

34.90

46.34

57.68

677.76

671.96

660.52

649.18

4.7067

4.6664

4.5869

4.5082

3206.3

3829.4

5042.2

6224.0

118.65

157.53

196.08

293.70

291.18

286.22

281.31

7.854 7.854 7.854

255.3336.1414.9

32 32.00

STD

XS

10

20 30 40

.312.375.500.625.688

31.37631.25031.00030.75030.624

31.06

37.26

49.48

61.60

67.68

773.19

766.99

754.77

742.64

736.57

5.3694

5.3263

5.2414

5.1572

5.1151

3898.9

4658.5

6138.6

7583.4

8298.3

105.59

126.66

168.21

209.43

230.08

335.05

332.36

327.06

321.81

319.18

8.378 8.378 8.378 8.378 8.378

243.7291.2383.7474.0518.6

34 34.00

STD

XS

10

20 30 40

.344.375.500.625.688

33.31233.25033.00032.75032.624

36.37

39.61

52.62

65.53

72.00

871.55

868.31

855.30

842.39

835.92

6.0524

6.0299

5.9396

5.8499

5.8050

5150.5

5599.3

7383.5

9127.6

9991.6

123.65

134.67

178.89

222.78

244.77

377.67

376.27

370.63

365.03

362.23

8.901 8.901 8.901 8.901 8.901

303.0329.4434.3536.9587.7

36 36.00

STD

XS

10

20 30 40

.312.375.500.625.750

35.37635.25035.00034.75034.500

34.98

41.97

55.76

69.46

83.06

982.90

975.91

962.11

948.42

934.82

6.8257

6.7771

6.6813

6.5862

6.4918

5569.5

6658.9

8786.2

10868.4

12906.1

118.92

142.68

189.57

236.13

282.35

425.92

422.89

416.91

417.22

405.09

9.425 9.425 9.425 9.425 9.425

309.4369.9488.1603.8717.0

Identification, wall thickness and weights are extracted from ANSI B 36.10 and B36.19. The notations STD, XS, and XXS indicate Standard, Extra Strong, and Double Extra Strong pipe respectively. Transverse internal area values listed in "square feet" also represent volume in cubic feet per foot of pipe length.

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Moody Friction Factor Calculator:


Table of Fluid Properties (Liquids and Gases)

Symbols: p (greek letter rho) = Density (units are mass/volume). The English (U.S. Customary Unit) for mass is the slug. The SI (metric) unit for mass is the kg. v (greek letter nu) = kinematic viscosity (units are length squared/time). If you're more familiar with dynamic viscosity µ (greek letter mu), then it may help to know that v = µ/p.

Fluid T (°F)

Density (slug/ft3)

v (ft2/s)

T (°C)

Density (kg/m3)

v (m2/s)

Liquids:

Water 70 1.936 1.05e-5 20 998.2 1.00e-6

Water 40 1.94 1.66e-5 5 1000 1.52e-6

Seawater 60 1.99 1.26e-5 16 1030 1.17e-6

SAE 30 oil 60 1.77 0.0045 16 912 4.2e-4

Gasoline 60 1.32 4.9e-6 16 680 4.6e-7

Mercury 68 26.3 1.25e-6 20 13600 1.15e-7

Gases (at standard atmospheric pressure, i.e. 1 atm):

Air 70 0.00233 1.64e-4 20 1.204 1.51e-5

Carbon Dioxide 68 0.00355 8.65e-5 20 1.83 8.03e-6

Nitrogen 68 0.00226 1.63e-4 20 1.16 1.52e-5

Helium 68 3.23e-4 1.27e-4 20 0.166 1.15e-4

The equations used in this program represent the Moody diagram which is the old-fashioned way of finding f. You may enter numbers in any units, so long as you are consistent. (L) means that the variable has units of length (e.g. meters). (L3/T) means that the variable has units of cubic length per time (e.g. m3/s). The Moody friction factor (f) is used in the Darcy-Weisbach major loss equation. Note that for laminar flow, f is independent of

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e. However, you must still enter an e for the program to run even though e is not used to compute f.

A more complicated equation which represents a slightly larger range of Reynolds numbers and e/D's is used in Design of Circular Liquid or Gas Pipes.

Minor Loss Coefficients, Hazen-Williams Coefficients, and Surface Roughness:

Table of Minor Loss Coefficients (K has no units)

Fitting K Fitting K

Valves: Elbows:

Globe, fully open 10 Regular 90°, flanged 0.3

Angle, fully open 2 Regular 90°, threaded 1.5

Gate, fully open 0.15 Long radius 90°, flanged 0.2

Gate 1/4 closed 0.26 Long radius 90°, threaded 0.7

Gate, 1/2 closed 2.1 Long radius 45°, threaded 0.2

Gate, 3/4 closed 17 Regular 45°, threaded 0.4

Swing check, forward flow 2

Swing check, backward flow infinity Tees:

Line flow, flanged 0.2

180° return bends: Line flow, threaded 0.9

Flanged 0.2 Branch flow, flanged 1.0

Threaded 1.5 Branch flow, threaded 2.0

Table of Hazen-Williams Coefficients (C has no units) To top of page

Material C Material C

Asbestos Cement 140 Copper 130-140

Brass 130-140 Galvanized iron 120

Brick sewer 100 Glass 140

Cast-Iron: Lead 130-140

New, unlined 130 Plastic 140-150

10 yr. old 107-113 Steel:

20 yr. old 89-100 Coal-tar enamel lined 145-150

30 yr. old 75-90 New unlined 140-150

40 yr. old 64-83 Riveted 110

Concrete/Concrete-lined:

Steel forms 140 Tin 130

Wooden forms 120 Vitrif. clay (good condition) 110-140

http://home.knpc.net/departments/maa/Operations/lc/pipeconversions.htm#Top of Page

Centrifugally spun 135 Wood stave (avg. condition) 120

Table of Surface Roughnesses To top of page

Material Surface Roughness, e

feet meters

PVC, plastic, glass 0.0 0.0

Commercial Steel or Wrought Iron 1.5e-4 4.5e-5

Galvanized Iron 5.0e-4 1.5e-4

Cast Iron 8.5e-4 2.6e-4

Asphalted Cast Iron 4.0e-4 1.2e-4

Riveted Steel 0.003 to 0.03 9.0e-4 to 9.0e-3

Drawn Tubing 5.0e-6 1.5e-6

Wood Stave 6.0e-4 to 3.0e-3 1.8e-4 to 9.0e-4

Concrete 0.001 to 0.01 3.0e-4 to 3.0e-3

Major Loss Calculation for Water in Pipes using Hazen-Williams Friction Loss Equation:


k is a unit conversion factor: k=1.318 for English units (feet and seconds). k=0.85 for SI units (meters and seconds) Rh=hydraulic radius=D/4 for circular pipe The Hazen-Williams method is only valid for water flowing at ordinary temperatures (about 40 to 75 oF). For other liquids or gases, the Darcy-Weisbach method should be used. Major loss (hf) is the energy (or head) loss (expressed in length units - think of it as energy per unit weight of fluid) due to friction between the moving fluid and the duct. It is also known as friction loss. The Darcy-Weisbach method is generally considered more accurate than the Hazen-Williams method. However, the Hazen-Williams method is very popular, especially among civil engineers, since its friction coefficient (C) is not a function of velocity or duct diameter. Hazen-Williams is simpler than

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Darcy-Weisbach for calculations where you are solving for flowrate, velocity, or diameter. More Discussion and References.

Major Loss Calculation for Fluid Flow using Darcy-Weisbach Friction Loss Equation:

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g = acceleration due to gravity = 32.174 ft/s2 = 9.806 m/s2 Major loss (hf) is the energy (or head) loss (expressed in length units - think of it as energy per unit weight of fluid) due to friction between the moving fluid and the duct. It is also known as friction loss. The Darcy-Weisbach method is generally considered more accurate than the Hazen-Williams method. Additionally, the Darcy-Weisbach method is valid for any liquid or gas; Hazen-Williams is only valid for water at ordinary temperatures (40 to 75 oF). The Hazen-Williams method is very popular, especially among civil engineers, since its friction coefficient (C) is not a function of velocity or duct diameter. Hazen-Williams is simpler than Darcy-Weisbach for calculations where you are solving for flowrate, velocity, or diameter. More Discussion and References.

Q=VA Flowrate Calculator


You may enter numbers in any units, so long as you are consistent. (L) means that the variable has units of length (e.g. meters). (L3/T) means that

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the variable has units of cubic length per time (e.g. m3/s). For further information, see Discussion and References.

Minor Loss Calculations for Fluid Flow

Minor loss (hm) is the energy (head) loss due to fittings (valves, elbows, etc.). hm has units of energy per unit weight of fluid which is the same as length.

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Static Pressure Calculation


Introduction Engineers and others often need a conversion between pressure and height of fluid column. For instance, if a swimming pool is 4 m deep, the pressure at the bottom can be computed. Or, if a fire hydrant must operate at a static pressure of 40 psi, the water column equivalent to 40 psi can be computed. Pressures can also be computed in terms of gas columns, so we have included some built-in values for gas density.

References All fluid mechanics textbooks discuss static pressure. A sampling of references follows.

Mays, L. W. editor. 1999. Hydraulic design handbook. McGraw-Hill Book Co.

Munson, B.R., D. F. Young, and T. H. Okiishi. 1998. Fundamentals of Fluid Mechanics. John Wiley and Sons, Inc. 3ed.

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Streeter, V. L., E. B. Wylie, and K. W. Bedford. 1998. Fluid Mechanics. WCB/McGraw-Hill. 9ed.

Energy Equation for Fluid Flow

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Calculation should be here.

The greek symbol "rho" is density (it is next to the g in the equation). g = acceleration due to gravity = 32.174 ft/s2 = 9.806 m/s2 If you know discharge and diameter (or area), go to Q=VA to compute velocity. The steady state energy equation for a fluid moving in a closed conduit is written between two locations at a distance L apart. The loss term hL accounts for all minor (valves, elbows, etc.) and major (pipe friction) losses. Often, one knows the flowrate and pipe diameter (or duct area) and is interested in determining the downstream pressure after a certain length of pipe. All of the losses would be solved for (minor and major), then the energy equation would be used to solve for the unknown pressure or whatever is unknown. Major losses can be computed using either the Darcy-Weisbach method or Hazen-Williams method. See Discussion and References for further information.

The energy equation is also valid for open channel flows.

Non-Circular Duct to Circular Pipe Conversions

Using these conversions, Design of Circular Liquid or Gas Pipes and Design of Circular Water Pipes can be used for non-circular ducts

Cross-sections:

Equations: For both geometries: Q=VA, Q=Flowrate, V=Velocity, A=Flow Area Rectangular: A=Area=WH. P=Perimeter=2(W+H). D=Hydraulic Diameter=4A/P Annular: A=Area=PI(D2

2-D12)/4. D=Hydraulic Diameter=D2-D1 References

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http://home.knpc.net/departments/maa/Operations/lc/flowrate.htm

http://home.knpc.net/departments/maa/Operations/lc/minorloss.htm





http://home.knpc.net/departments/maa/Operations/lc/DarcyWeisbach.htm

http://home.knpc.net/departments/maa/Operations/lc/HazenWilliamsDesign.htm


http://home.knpc.net/departments/maa/Operations/lc/literature.htm#References


Enter all values in consistent units: If you enter diameter in inches, then width and height will be output in inches. If you enter width in cm, then you must also enter height in cm; and diameter will be output in cm. If velocity is entered in cm/min, then flowrate will be cm3/min.

Using this page: If you have a non-circular duct, the cross-sectional geometry can be converted to an hydraulic (equivalent) diameter using this page. Then, velocity (not flowrate) can be computed using a) Design of Circular Liquid or Gas Pipes or b) Design of Circular Water Pipes based on the hydraulic diameter. However, (a) and (b) will not compute the correct flowrate since they compute area based on a circular cross-section. Therefore, you must multiply the velocity by the actual area of the duct. The actual area is computed on this page.

If you wish to determine the cross-sectional geometry of a non-circular duct, you can use a) Design of Circular Liquid or Gas Pipes or b) Design of Circular Water Pipes if you know the velocity, not the flowrate (flowrate can only be converted to velocity if the actual duct area is used, not the area based on hydraulic diameter). From the hydraulic diameter output from either (a) or (b), this page can be used to determine the geometry of the duct.

Manning's Equation Calculator


The Manning Equation is the most commonly used equation to analyze open channel flows. It is a semi-empirical equation for simulating water flows in channels and culverts where the water is open to the atmosphere, i.e. not flowing under pressure, and was first presented in 1889 by Robert Manning. The channel can be any shape - circular, rectangular, triangular, etc. The units in the Manning equation appear to be inconsistent; however, the value k has hidden units in it to make the equation consistent. The Manning Equation was developed for uniform steady state flow (see Discussion and References for Open Channel Flow). S is the slope of the energy grade line and S=hf/L where hf is energy (head) loss and L is the length of the channel or reach. For uniform steady flows, the energy grade line = the slope of the water surface = the slope of the bottom of the channel.

The product A/P is also known as the hydraulic radius, Rh.

Manning's n Coefficients for Open Channel Flow

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Material Manning n Material Manning n

Natural Streams Excavated Earth Channels

Clean and Straight 0.030 Clean 0.022

Major Rivers 0.035 Gravelly 0.025

Sluggish with Deep Pools 0.040 Weedy 0.030

Stony, Cobbles 0.035

Metals Floodplains

Brass 0.011 Pasture, Farmland 0.035

Cast Iron 0.013 Light Brush 0.050

Smooth Steel 0.012 Heavy Brush 0.075

Corrugated Metal 0.022 Trees 0.15

Non-Metals

Glass 0.010 Finished Concrete 0.012

Clay Tile 0.014 Unfinished Concrete 0.014

Brickwork 0.015 Gravel 0.029

Asphalt 0.016 Earth 0.025

Masonry 0.025 Planed Wood 0.012

Unplaned Wood 0.013

Corrugated Polyethylene (PE) with smooth inner walls a,b 0.009-0.015 Corrugated Polyethylene (PE) with corrugated inner walls c 0.018-0.025 Polyvinyl Chloride (PVC) with smooth inner walls d,e 0.009-0.011

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