Aurora Pump School 101

(1)

Centrifugal

Pumps 101

School

(2)

Centrifugal Pumps 101

Part 1 Basic Pump Design & Parts Terminology

The following presentation is not intended to cover every type of centrifugal pump, but the basic components from pump type to pump type are typically the same. You will learn about the two most popular pump designs, and the major components which make up these pumps. There is still much to be learned about pumps, and this is meant to be a beginners course only. It is recommended that you visit a pump factory within the first few months of working in this industry to get a better understanding of what you are learning today.

(3)

Basic Pump Design and Parts Breakdown

For the purposes in this study, we will be

including only split case and end suction pumps.

We will give a basic description of each, show an

actual picture of each type pump, show a brochure

type cutaway, and a parts breakdown from a repair

manual. Although there are many small parts,

nuts, bolts, gaskets, etc….. used to build a pump,

we will only be concentrating on the major

components that are generally common

throughout the industry, so you can have a basic

understanding of pump design.

(4)

End Suction Pumps

–

_–

Frame MountedFrame Mounted

z End suction pumps are generally broken down into two categories, close coupled and frame mounted. The term end suction implies that water enters the pump from the end of the unit as you will see in our diagrams. Our first example will be of the frame mounted design. Frame mounted implies that the pump has a bearing frame and shaft of it’s own, separate from the driver or motor. The motor also has a shaft and bearing frame of it’s own. A coupling is the component that attaches to both the pump and motor shaft, and makes a connection between the two. The pump and motor shafts have to be lined up almost perfectly for the unit to function properly. This is referred to as the pump or coupling alignment. The major components of a frame mounted end suction pump are the casing, impeller, casing wear ring, mechanical seal or packing, shaft, shaft sleeve, bracket, bearing frame and bearings. The impeller and casing are the two most expensive items in the pump, and the casing houses the impeller and includes the suction and discharge ports The casing wear ring is an inexpensive replaceable item, and protects the casing from wear in the most vulnerable area. The shaft sleeve fits over the shaft and protects the more expensive shaft from wear in the same way. All pumps have a suction and discharge. The suction is where the water enters the pump, and the discharge is where the water leaves the pump. Most often, the suction is larger than the discharge, occasionally they are the same size. Study the major components of these pumps on the following slides.

(5)

Frame Mounted End Suction Pump

Coupling and Coupling Guard. The actual coupling is under the

guard and unseen in this picture. Frame Mounted Motor

(6)

As the name implies, the pump takes suction from the end. Bearing frame Pump shaft Bearings Bracket

Mechanical seal Shaft sleeve

Casing

Impeller Case wear ring

(7)

Item and number Mechanical seal #27 Shaft sleeve #25 Impeller #11 Casing #6 Case ring #7 Bracket #35 Bearing Frame #57 Bearings #54 & 53 Shaft #55

(8)

End Suction Pumps

–

_–

Close CoupledClose Coupled

Close coupled end suction pumps are similar to frame mounted end suction pumps. The impeller and casing are usually the same. The big difference is in the motor and bearing frame of the pump. A close coupled pump does not require a coupling since the pump does not have a shaft and bearings of it’s own. The motor is built with an extended shaft, and the bracket makes the connection between the pump and motor. The impeller is actually attached to the motor shaft. When repairs have to be made to end suction pumps, the adapter and casing are unbolted from each other, and the casing remains attached to the system piping, while the rest of the pump is pulled back away from the casing. This is referred to as a back pullout design. In the case of a frame mounted end suction pump, the motor usually has to be unbolted from the base to allow the pump frame to be pulled away from the casing. Study the following slides on close coupled end suction pumps.

(9)

Close Coupled End

Suction Pump Bracket

Close Coupled Motor

(10)

Motor Bracket _Casing

Impeller

Case ring Mechanical seal

(11)

Motor Item & Number

Bracket #18 Mech. Seal #11 Impeller # 9 Case ring # 5 Casing #4

(12)

Horizontal Split Case Pumps

Horizontal split case pumps are just as the name implies. The case is split in half on a horizontal plane. Because the pump shaft exits each side of the casing, there are two shaft sleeves, two case rings, and two sets of mechanical seals or packing. Both sides of the pump shaft are supported by bearings, and this decreases shaft deflection and increases pump life. The suction flange is on one side of the pump, and the discharge flange is on the opposite side of the pump. Split case pumps do not come with a close coupled option, and since the casing houses both bearing assemblies, there is also no bracket needed. All split case pumps are mounted on a base plate and connected to the motor with some type of coupling. When working on a split case pump, the upper half of the casing is removed, and the entire rotating assembly, (all components attached to the shaft) can be removed and repaired without disturbing the piping or motor. Only the coupling has to be disconnected. Please study the following slides on horizontal split case pumps.

(13)

Motor Coupling Guard

Horizontal Split Case Pump

(14)

Pump Shaft

Lower Half Pump Casing

Upper Half Pump Casing _Impeller

Packing**

Mechanical Seal**

Shaft Sleeves

Case Rings Bearings

For display purposes only, this pump has packing on one side and

(15)

Upper Half Pump Casing Mechanical Seal** Shaft Sleeves Bearings Case Rings Impeller Packing** Pump Shaft

Lower Half Pump Casing

For display purposes only, this pump has packing on one side and mechanical seals on the other side. Normally, both sides would be the same.

(16)

Items & Number Upper case #8 Lower case #69 Shaft sleeve #57 Shaft sleeve #64 Bearing #38 Bearing #47 Shaft #65 Impeller #59 Case ring #28 Case ring #28

(17)

Mechanical Seals

z Mechanical seals are one method used to seal the pump

shaft where it exits the pump casing. Mechanical seals are the preferred method of sealing a pump when the liquid being pumped is clean and clear, because they typically do not allow any leakage from the pump. The rotating face turns on the surface of a stationary face, along with gaskets, o-rings and bellows to form a seal. The following slide shows a two stage pump (two impellers) with mechanical seals on one side and packing on the other side. This is done for display purposes only. For actual applications, pumps are always built with the same method of sealing on both sides of the pump. As an exercise on what we have already studied, can you name the first 5 numbered parts on the next slide as well?

(18)

Stationary Seat

Rotating face, and bellows.

(19)

Packing

z Packing is another method used to seal the pump shaft

where it exits the pump. Some liquid must leak past the packing rings to dissipate the heat generated by friction between the stationary packing and the rotating shaft sleeve underneath the packing rings. This leakage must be piped to drain, or collected in a sump to be pumped out of the pump room. Because of this necessary leakage, packing is not the preferred method of sealing a pump. However, packing is generally not subject to a catastrophic failure, and for this reason, packing is the only accepted method of sealing UL/FM approved fire pumps.

(20)

(21)

End Suction VS Split Case

As you can see, end suction pumps are smaller, have less parts, and because they are less expensive, they are often used in place of split case pumps on small to medium applications. There are two major differences in these two types of pumps, and the split case pump is the superior design in both cases. An end suction pump has what is referred to as an overhung impeller. In other words, both shaft bearings are on one side of the impeller, so the pump shaft is not supported past the last bearing where the impeller is attached. A split case pump has a bearing on each side of the impeller, so the shaft is supported from both ends. The split case design gives better shaft support which leads to longer pump life. The end suction pump also has a single suction impeller (water enters only from one side), while the split case pump has a double suction impeller. Water entering the impeller from one side only creates a hydraulic thrust imbalance, and this causes increased thrust and shaft deflection (bending) and increases wear on bearings and seals. Split case pumps have double suction impellers, and so the hydraulic thrust is balanced.

(22)

Centrifugal Pumps 101

Part 2 Self Taught School of Basic Hydraulics

The following presentation is not intended to be a complete study of pump hydraulics. Although this is a basic beginners course, much of the information needed to be successful in the pump industry is included in this study. However, there will be topics which are not fully discussed, as they would not be considered beginners material. Once you have a basic understanding of pump hydraulics, it is recommended that you obtain a copy of “Cameron Hydraulic Data” and “Hydraulic Institute Standards” for further study.

(23)

This study will show the basic design criteria used

by engineers to make pump selections, review

basic pump terms and formulas, and give

examples on proper and improper pump selection.

You will learn how to convert PSI (pounds per

square inch) to feet of head, how to calculate the

horsepower required to operate any given pump,

and how to read pump curves and properly make

pump selections. You will also be given examples

throughout the program on common mistakes that

will cause premature pump failure.

(24)

What we are trying to

accomplish!

z

In most pumping applications, there are

generally two factors an engineer is trying

to overcome, and a final goal he is trying to

accomplish.

(25)

Friction Loss

z The first factor present in every pump application that we

are trying to overcome is friction loss. Simply stated, this is the drag created when flowing water comes in contact with the inside wall of the piping, valves and fittings of a system. Friction loss can be calculated with formulas, friction loss charts or computer programs. As you can see by the friction loss chart on the next slide, increased flow in the same size pipe will increase friction loss. You will also notice that the friction loss given is per 100’ of pipe, so more piping and valves in a system will also increase overall friction loss. These same types of charts are available for valves and fittings. The second slide depicts a system with 24’ of friction losses in the discharge piping.

(26)

At 100 GPM, the friction loss is only .094 feet per 100 feet of 6” pipe. At 650 GPM, the friction loss is now 3.38 feet per 100 feet of 6” pipe.

(27)

(28)

Elevation or Static Loss

z The second factor present in most pump applications that we

are trying to overcome is Elevation loss, or Static loss. Simply stated, Static loss is the difference in elevation from point A to point B in a piping layout. In the example on the next slide, the Elevation or Static loss is 35 feet. Friction losses and static losses are both calculated in feet of head, which is an expression of pressure, and they can also be expressed as PSI. Formulas will be given later in this presentation to convert PSI to feet of head.

(29)

(30)

Final Goal

z

After we have overcome the friction loss and any

static loss in the system, there is usually a fixture or

some piece of equipment which requires a set water

pressure and flow to operate properly. In the

example on the next slide, we have a chiller which

requires 600 gallons per minute, and 50 PSI to

operate properly. This is a basic example. Most

applications require different flow rates in the same

system. In this case, the engineer is then going to

do calculations using system head curves to

determine pump selection. Please refer to the tool

kit covering system head curves for further study.

(31)

(32)

Tying It All Together

z In the example on the next slide, we have a chiller that

requires 600 GPM @ 50 PSI (115’) to operate properly. We have friction losses of 24’ (aprox. 10 PSI), and we have static losses of 35’ (aprox. 15 PSI). Adding these three together, the total system requirement is therefore 174’ or aprox. 75 PSI. We also have a suction pressure of 32 PSI (aprox. 74’) entering the pump, so we can deduct that from the required 174’. The end result is that our pump requirement is 600 GPM @ 100’ or aprox. 43 PSI. This is the process an engineer goes through to pick a pump selection on any given job. After you have gone through this entire presentation, it is advisable to study it once more. Many of the examples will make more sense after you have a basic understanding of pump terms and formulas.

(33)

(34)

Basic Centrifugal Pump Terms

and Formulas

z

The following slides will list common pump

terms and formulas used in the pumping

industry. Each will be explained and

examples will be given. It is important that

you take the time to memorize these terms

and formulas as you will run across most of

these on a daily basis.

(35)

GPM = GALLONS PER MINUTE

z Simply stated, fluid being pumped is measured in gallons per

minute. Two other terms used to express GPM are flow and capacity. Pump curves as seen on the next slide are shown with GPM measured along the bottom of the curve, which is industry standard. Since most pumps are marketed world wide, you will notice the flow is also measured in L/S (liters

per second) and M3_{/HR (cubic meters per hour). In the}

United States, pump flow is almost always expressed in GPM. In the example on the next slide, the design point marked with a red dot shows a pump operating at 600 GPM.

(36)

(37)

PSI = Pounds per Square Inch

z PSI is one way of expressing pressure developed by a

pump. Because most pressure gauges used on pumps are calibrated using a PSI scale, this is a common term used in the field. However, most pump curves express pressure in feet of head. This is because pressure will vary with different liquids being pumped, while feet of head remains constant. This will be explained further under Specific Gravity. In the example on the next slide, pressure is noted on the left side of the curve which is industry standard. You will also notice the pressure on the curve expressed as meters. Once again, this is for pumps being sold overseas. In the example given on the next slide, the design point shows the pump operating at 100 feet of head.

(38)

(39)

Head

z

Head, or feet of head is the most common way to

express pressure generated by a pump. Most pump

curves use feet of head to express pressure. A

formula to convert PSI to feet of head will be given

later in the presentation. In the example on the next

slide, the pump design point selected shows a pump

operating at 600 GPM @ 80 feet of head (80’).

(40)

(41)

TDH = Total Dynamic Head

z Total dynamic head is often confused with feet of head.

Technically speaking, TDH is the total discharge head, minus the total suction head, and implies that the consulting engineer has completed all necessary system calculations. As an example, if the total pump discharge pressure required is 140’ of head, and the suction pressure is 60’ of head, the pump TDH is 80’. This 80’ should include the systems friction and elevation losses on the discharge side of the pump, along with the requirements of the equipment or fixture in the system, and the suction pressure should already have been subtracted from this number. This 80’, along with the GPM will comprise the design point of the pump.

(42)

Conditions or Conditions of Service

z When you hear the term conditions, conditions of service, or

design point being referred to, reference is being made to the GPM and TDH required on any given pumping application. In the previous examples we have given, the conditions of service were 600 GPM @ 100’ of head, and 600 GPM @ 80’ of head. Can you read the pump curve on the following slide and determine what the conditions of service are for the point selected? ……… The correct answer can be found on the slide after the pump curve. If you did not pick the correct conditions, go back and review the last seven slides.

(43)

(44)

Impeller Diameter

(The correct answer on the previous curve was 1200 GPM @ 95’)

The impeller is the heart of the pump, and as water enters and passes through the impeller, pressure is increased. Increasing the impeller diameter increases the centrifugal force being exerted on the water, and so increasing impeller diameter will also increase the pressure created by the pump. The following slide points out the different flow characteristics of various size impellers within the same pump. Only a few impeller diameters are shown, but the impeller can be trimmed to many diameters in between those shown. The actual impeller diameters are shown in inches on the left side of the curve near the beginning of the pump curve.

(45)

Different Impeller Trims

(46)

Curve Shape

You will notice that as the pump curve moves to

the right, (flowing more water) the curve is also

slowly dropping (less pressure). The reason for

this is that as the flow increases, the same water

remains inside the impeller for a shorter period of

time, and the centrifugal force has less time to

impact the passing water. You will notice in the

future that some curves are flatter than others, and

some curves drop off quicker as you flow more

water. The reason for this difference is in the

internal impeller design. The following slide

points out the natural curve shape of a centrifugal

pump from beginning to end for a 13 inch

impeller.

(47)

(48)

RPM = Revolutions Per Minute

z Simply stated, revolutions per minute are the number of times a

pump shaft will fully rotate in one minute. Common electric motor nominal speeds are 3500, 1750, 1150, and 870 RPM. Increasing pump RPM’s will increase GPM and feet of head. A small pump turning at a faster speed can often meet or exceed the conditions of a larger, slower turning pump. A standard motor will turn at one speed, while a motor controlled by a VFD (variable frequency drive) will change speed to match conditions of service. Engine driven pumps can operate at many different RPM’s determined by the engine governor. In general, higher RPM pumps will require more maintenance than slower turning pump. All of the normal wear items such as bearings, case rings and packing or mechanical seals will wear out quicker on the faster turning pumps.

(49)

Efficiency

z A pumps efficiency is the ratio of input power to output

power of the pump, and is expressed as a percentage. (%) Many factors affect a pumps efficiency which include, internal friction loss as water passes through the casing and impeller, and mechanical drag created by the bearings, mechanical seals or packing, and wear rings. A pumps efficiency is shown on the pump curve most commonly in one of two ways shown on the next two slides. In the two examples given, the same design point is selected using two different pumps, showing pump efficiency’s of 80% and one of 68%. In today’s world of energy conservation, it is usually very important to have a pump selection with a high efficiency rating. Many specifications will call out a minimum efficiency required to bid on a given project.

(50)

Every point within these boundary lines will have an efficiency of 80%

Other efficiency curves as well.

(51)

This is the other common way to show pump efficiency on a curve. The green curve is the efficiency curve, and the corresponding values are shown on the right side of the graph. This pump shows an efficiency of 68%

(52)

BEP = Best Efficiency Point

z

The best efficiency point on any given curve is found

at the GPM where the efficiency rating is the highest.

The next slide shows a pump curve that has a BEP

between approximately 2000 and 2500 GPM with a

full diameter impeller. Selecting a pump at or near

it’s BEP is in the customers best interest, and will

probably increase the service life of the pump. If you

cannot select a pump at the BEP, it is generally better

to select a pump with a point that is slightly to the left

of the BEP rather than to the right of the BEP. The

reason for this is that we are trying to avoid operating

near a runout condition, which will be discussed

shortly.

(53)

(54)

Shutoff

z Shutoff, shutoff pressure or churn pressure, is the pressure developed when the pump is running, but not flowing any water. (0 GPM) This could be due to a closed valve, or a pump running in a system that is not currently requiring any water. The shutoff pressure for any given pump can be found by following the pump curve all the way to the left side of the curve. If the design point you have selected is between impeller diameters, follow an imaginary line to the left side of the curve. The curve on the following slide shows a shutoff head of approximately 114’. Centrifugal pumps are very popular because shutoff pressures usually do not exceed 20% of the design pressure. This is important because it reduces the possibility of over pressurizing a piping system. Running pumps close to shutoff pressure for extended periods of time will increase shaft deflection and reduce pump service life. Running a pump without a casing relief valve at shutoff for more than a few minutes can cause serious heat build up and damage or destroy the pump. A casing relief valve is installed on the discharge side of the pump and piped to a drain, and it is set to open at a pressure slightly less than shutoff pressure to prevent the pump from operating at a zero flow condition.

(55)

Shutoff

(56)

Runout

Runout is a term used when a pump is operating to

the far right of the pump curve. Efficiency is

reduced, NPSHr is increased, and shaft deflection

increases. For these reasons, it is not advised to

operate a pump at this point on the curve. Follow the

design point on the pump curve on the next slide to

a runout condition. This would be near 1000 GPM.

(57)

Runout

(58)

Manufacturer’s Warning

z

Many pump manufacturer’s recommend that their

product not be operated at certain points on the

pump curve. This can be done in several ways,

either by notes on the pump curve, or boundary

curves. The example on the next slide shows

boundary curves, and it is recommended that the

pump not operate to the left of the curve near

shutoff, or to the right of the curve near runout.

Running a pump outside of the recommended

operating range will usually void the

manufacturer’s warranty.

(59)

b

Boundary line warning curves

(60)

BHP = Brake Horsepower

z BHP or brake horsepower is the power required to operate a

pump at any given point on the pump curve. The curve on the following slide shows a pump with a brake horsepower requirement of approximately 45 at design point A. You will notice that design point A falls half way between the 40 and 50 horse power curves. Most pump curves show horse power curves, and this is a good method of estimating horse power requirements at a glance. Later in the presentation, a formula will be given to calculate brake horsepower. It is important to memorize this formula.

(61)

Point A

(62)

Max. BHP = Maximum Brake Horsepower

z Max. BHP is the maximum horsepower required to run any

given pump at any point along the pump curve. In the example on the next slide, the BHP at design point A is approximately 45. The max. BHP for this same pump is approximately 76. To find this point, follow the pump curve all the way to the right just as we did when looking at pump runout. When sizing a motor, it is important to take the max. BHP into consideration since the pump, once installed into a system, will very likely run at a number of different points along the curve. If a 50 HP motor were to be sold with this pump, there is a good chance the motor could be overloaded sometime in the future if the pump operates to the right side of the curve.

(63)

Point A Max. BHP

(64)

Motor Service Factor (S.F.)

z The motor service factor can also be looked at as a safety factor. Most motors built today have a 1.15 S.F. What this means to you is that the motor has a 15% built in safety factor. As an example, a 100 HP motor with a 1.15 S.F. is actually capable of 115 ( 100 X 1.15 ) horse power output. Some engineers will allow you to use the service factor, others will not. When you hear the term non overloading throughout the curve, this means you cannot use the S.F. when making your motor selection. Running a motor into the service factor does increase heat developed by the motor, and will shorten the lifespan of the motor. It is best to use the S.F. only in situations where the pump will occasionally run to this point on the curve. As an example, if you have a pump with a BHP of 45 and a Max. BHP of 51, you can probably be safe by using a 50 HP motor instead of having to increase to a 60 HP motor. Not all motors have a 1.15 service factor, so check before making a mistake. A motor with a 1.0 S.F. basically has no safety factor. A 100 HP motor with a 1.0 S.F. is only good up to 100 HP.

(65)

Motor Enclosures

Motor enclosures mainly refer to the type of environment a motor is suitable to be used in. Common examples would be:

ODP – open drip proof – to be used indoors and usually has a service factor of 1.15

TEFC – totally enclosed fan cooled – can be used in

outdoor applications and commonly has a service factor of 1.0

Explosion Proof – normally used in areas where hazardous or combustible vapor conditions can exist and usually has a service factor of 1.0

(66)

NPSH = Net Positive Suction Head

z

NPSH is probably the least understood and most

confusing concept in pump application. Simply

stated, we are pumping liquid with the type of

pumps we sell, and NPSH relates to the fact that

we need to make sure the liquid stays in liquid

form, and does not begin to become a mixture of

liquid and gas. This tool kit contains an in depth

discussion on NPSH and it is recommended that

you study this information as well.

(67)

NPSH

-

_-

Continued

_Continued

z NPSHa refers to the amount of net positive suction head that is available in any particular application.

z NPSHr refers to the amount of net positive suction head required by any given pump to operate properly. The NPSHa must always exceed the NPSHr if the liquid being pumped is to stay in liquid form.

z This tool kit contains a separate, in depth discussion on NPSH. Please refer to that section for more information.

z Most pump curves show NPSH requirements. Increasing GPM increases NPSH requirements. Most applications that take suction from a city main have more than adequate NPSHa. Hot water applications and cooling tower applications are two situations where NPSH can be critical.

z In the example on the next slide, the design point shown has an NPSHr of about 12’. To be safe, this pump should not be used in an application that has less than about 16’ of NPSHa. As we have discussed earlier, the pump has a good chance of operating to the right or left of the design point, and you should always have a margin of safety included in your selection.

(68)

NPSH Curves

Design Point

(69)

Vapor Pressure

z Vapor pressure is the amount of pressure required to keep

any given liquid in liquid form. As you can see in the chart on the next slide, the vapor pressure increases as the temperature of the liquid increases. 212 degrees Fahrenheit is the boiling point of water at sea level, since the vapor pressure at 212 degrees equals that of the atmospheric pressure at sea level or 14.7 psi absolute (psia). You will notice that at 300 degrees, the vapor pressure is 67 psia. What this means is that if you want to keep water in liquid form at 300 degrees F. you would need at least 67 absolute pounds of pressure being exerted on the water. The only way this is going to happen would be in a pressurized vessel.

(70)

Vapor Pressure of Water at Various Temperatures Vapor pressure of water at 212 degrees F. Vapor pressure of water at 300 degrees F.

(71)

Cavitation

Cavitation is the condition of pumping a mixture of liquid and air, either from entrained air, vortexing or an application with insufficient NPSH. Typical cavitation, due to a lack of proper NPSHa, will outwardly sound as if rocks are passing

through the pump. Cavitation will eventually destroy a

pump. In severe cases, an impeller can be destroyed in a week or two. Entrained air is air bubbles that are trapped in the water. This is usually caused by turbulent water that does not have time to let the bubbles rise to the surface and pop. We have all seen an example of a vortex in the bathtub at our homes. As the water drains, a tornado looking whirlpool appears. This allows air down into the drain in much the same way it would happen with a pump.

(72)

Specific Gravity

z The specific gravity of a liquid is it’s density or weight

compared to that of water at ambient temperature. Room temperature water has a specific gravity of about 1.0 The chart on the next slide shows the S.G. of various liquids. The specific gravity does affect pressure expressed in feet of head, along with the horsepower requirements of a pump. While it is important to know this, most applications involve water at room temperature and require no adjustments for S.G. From the chart on the next slide, you will notice the first liquid is over 150% heavier than water. If placed in water, it would sink. You will also notice the last liquid is over 20% lighter than water. If placed in water, it would float.

(73)

1.594 0.79 Over 150% heavier than water. Over 20% lighter than water.

(74)

Shaft Deflection

z

Shaft deflection is the amount a pump shaft will

bend due to radial, hydraulic or mechanical loads.

Running a pump to the far left near shutoff, or to the

far right near runout, will increase pump shaft

deflection. Increased shaft deflection will cause

premature bearing, and seal failure, and can also

cause pump shaft breakage. Undersized pump

shafts and running a pump at speeds higher than it

was designed for will also cause increased shaft

deflection and reduced pump life. The curve on the

next slide reminds you of the shutoff and runout

conditions.

(75)

Runout

Design point

(76)

Pump Formulas

z

Converting PSI to Feet of Head.

PSI X 2.31 = Feet of Head

Specific Gravity (S.G.)

When pumping fresh water of ambient temperature, the specific gravity value will be approx. 1.0 and can be omitted from the equation. Pumping hot water or liquid other than water are notable exceptions. As you can see from the examples below, the hotter lighter liquid requires more feet of head to equal the same 100 PSI.

Example: 80 degree F. water at 100 PSI 100 X 2.31 = 231’ of head 1.0

Same 100 PSI at 240 degree F. water 100 X 2.31 = 244’ of head .9464

(77)

Formulas Continued

Formula to calculate BHP.

GPM X TDH X S.G. = BHP

3960 X pump efficiency (pump efficiency is found on the pump curve)

Under normal conditions, pumping fresh water at ambient temperature, the S.G. is aprox. 1.0 and can be left off of the equation.

Example: 1000 GPM @ 200’ @ 78% efficiency pumping 80 degree water. 1000 X 200 X 1.0 = 200,000 = 64.75 BHP

3960 X .78 3088.8

If we use the same example as above, but use 240 degree water, the

results change as follows because the S.G. of 240 degree water is .9464. 1000 X 200 X .9464 = 189,280 = 61.28 BHP

(78)

Formulas Continued

z Velocity through a piping system is expressed in feet per second. As velocity increases, friction loss increases, and hydraulic noise through the piping will increase as well. In general, it is good practice to see velocities at or below 8 feet per second. Below is the formula to calculate velocity through a pipe.

z Velocity = .4085 X GPM (where d is the inside diameter of the pipe in inches)

d 2

Example: Pumping 700 GPM through a 8” pipe will have the following velocity.

.4085 X 700 = 285.95 = 4.467 feet per second 8 X 8 64

(79)

Formulas Continued

z The Affinity Laws are formulas showing the effects of changing the pump speed or impeller diameter of a given pump. Increasing the speed or impeller diameter, increases the conditions and BHP. Decreasing the speed or impeller diameter decreases the conditions and BHP. The example we are giving is with RPM, but you can use two different impeller diameters and determine the new operating conditions as well. 1 2 RPM₂ relates to GPM RPM₂ = relates to TDH RPM₁ RPM₁ 3 RPM₂ = relates to BHP RPM₁

(80)

Affinity Laws Laws Continued

z Example: We are pumping 1000 GPM @ 200’ @ 3500 RPM with a BHP of 64.75 To find out the new set of conditions with the pump operating at 2800 RPM, we apply the affinity laws as follows:

2800 1 _{= .80 relates to GPM 2800} 2 _{= .64 relates to TDH}

3500 3500

2800 3 _{= .512 relates to BHP}

3500

Using the example above, the new conditions of service at the new pump speed are:

1000 GPM X .80 = 800 GPM 200’ X .64 = 128’ 64.75 BHP X .512 = 33.15 BHP

(81)

POP QUIZ

It’s time to apply what you should have learned

while studying this program. A potential customer

calls your office with the following quote request.

“Please quote me a split case pump capable of

pumping 1100 GPM @ 65 PSI. I will be

pumping 80 degree fresh water and have 20’ of

NPSH available”. --- Look over the

following four pump curves, make the best

selection based on the information given and

determine what size motor you will use. The

correct answer along with explanations will

follow the four pump curves.

(82)

(83)

(84)

(85)

(86)

z First things first. Hopefully everyone successfully converted the 65 PSI to 150 feet of head. The pump curves selected for your review were selected so that the best pump would not be overly obvious. If you did not pick the best pump, don’t be discouraged. You have been exposed to a large amount of

information, and it will take some time for everything to sink in and make sense. Please review all of the choices below to get the most out of this exercise.

z 8 X 8 X 11B - If this was your choice, this selection is a bit to the left side of the curve, but not too bad. The efficiency is fairly good, and the NPSHr is OK. But since this pump starts off so far to the left of the curve, the horse power at runout exceeds 75 HP, and so a 100 HP motor would be needed to insure a non overloading condition. Therefore, this selection would not be economically feasible.

z 4 X 5 X 11C - If this was your choice, this selection falls in a great spot on the pump curve. The

efficiency is good, and we are at the BEP. At runout, we still only require a 60 HP motor, which looks good. However, we only have 19’ of NPSH available, which is almost exactly what the pump needs to operate cavitation free. If any calculation is only slightly off, or if the pump operates any more to the right of the curve at all, this pump will certainly cavitate. Therefore, this is not a good selection.

z 5 X 6 X 15 - If this was your choice, this selection falls in a great spot on the pump curve. The

efficiency is good and we are almost at the BEP. At runout, we are borderline between using a 60HP and 75HP motor. With a 1.15 service factor motor, it should be safe to use the 60 HP motor. The NPSHr of the pump is about 12 or 13, and we can run out to about 1500 GPM on the pump curve before NPSH becomes an issue. Of the four pumps given to choose from, this is the best selection.

z 4 X 5 X 10B – If this was your choice, this selection falls in a poor spot on the pump curve. We are too far to the right of the curve near runout, the horse power requirement is good at 60 HP, but the biggest problem is that the NPSHr for this pump is over 35’. Not only will this pump certainly cavitate, it probably won’t last more than a week or two.