Modifications, Edits and Additions to the CD
Corrected and Edited papers
A009. Affnity laws for centrifugal pumps
C012-1 Carbonic acid
D035. Dynamic balance
M029 Motor selection
N001 NPSHA in USCS units
P033. Pressure head
P015 Piping recommendations
P042 Some common misconceptions about centrifugal pumps
S039 Shaft bending, the L3D4 formula
S072 Specific speed
SA005-1 Temperature limits for elastomers
V026 Vortexing Lliquid
W007 Wear ring clearance
F040 Seal FAQ
C075 Iron corrosion
C076 Corrosion problems with submersible pumps
S116 Surging pump
N001. NPSHA. CALCULATING NET POSITIVE SUCTION HEAD AVAILABLE IN USCS (INCH) UNITS
Change the eighth paragraph from the botom to:
" NPSHA (net positive suction head available) = 34 + 5 - 22.7 - 0.62 - 2.34 = 13.9 feet. This is enough to stop cavitation also.
P033. PRESSURE HEAD - The last formula in the article shows a (x 2) It should read
S039. SHAFT BENDING, LEARNING THE L3/D4 FORMULA
Under the paragraph CALCULATING SHAFT DEFLECTION Change the first formula to:
A009. AFFINITY LAWS FOR CENTRIFUGAL PUMPS
There are occasions when you might want to vary the amount of fluid you are pumping or change the discharge head of a centrifugal pump. There are four ways you could do this:
- Change the speed of the pump.
- Change the diameter of the impeller. Replace it with a larger impeller, or cut down the one you have.
- Purchase a different pump with the operating curve you need.
- Valve or orifice the discharge of the pump to get the capacity or head you need. Unfortunately this can cause the pump to operate off of its best efficiency point (BEP).
Of the four methods the first two are the only sensible ones unless you are prepared to buy a new pump. In the following paragraphs we will learn what happens when we change either the pump speed or impeller diameter, and as you would guess, other characteristics of the pump are going to change along with the speed or diameter.
To determine what is going to happen, you begin by taking the new speed or impeller diameter and divide it by the old speed or impeller diameter. Since changing either one will have approximately the same affect,
I will be referring only to changing the speed in this part of the discussion.
As an example:
The capacity or amount of fluid you are pumping will vary directly with this number.
100 Gallons per minute x 2.0 = 200 Gallons per minute
50 Cubic meters per hour x 0,5 = 25 Cubic meters per hour
The head produced by the pump varies by the square of the number.
A 50 foot head x 4 (22) = 200 foot head
A 20 meter head x 0,25 (0,52) = 5 meter head
The horsepower required changes by the cube of the number
A 9 Horsepower motor was required to drive the pump at 1750 rpm. How many horsepower is required now that you are going to 3500 rpm?
9 x 8 (23) = 72 Horsepower is now required.
Likewise if a 12-kilowatt motor was required at 3000 rpm. and you decreased the speed to 1500 the new kilowatts required would be:
12 x 0,125 (0,53) = 1,5 kilowatts required for the lower rpm.
The following relationships are not exact, but they give you an idea of how speed and impeller diameter affects other pump functions.
The net positive suction head required by the pump manufacturer (NPSHR) varies by the square of the number.
A 3 meter NPSHR x 4 (22) = 12 meter NPSHR
10 foot NPSHR x 0,25 (0,52) = 2.5 foot NPSHR
The amount of shaft run out (deflection) varies by the square of the number
As an example: If you put a dial indicator on the shaft and noticed that the total run out at 1750 rpm. was 0.005 inches, then at 3500 rpm the run out would be 0.005" x 4 or 0.020 inches.
Likewise if you had 0,07 mm. run out at 2900 rpm. and you slowed that shaft down to 1450 rpm the run out would decrease to 0,07 mm x 0,25 or 0,018 mm.
The amount of friction loss in the piping varies by 90% of the square of the number. Fittings and accessories vary by the square of the number.
As an example: If the system head loss was calculated or measured at 65 meters, at 1450 rpm. the loss at 2900 rpm. would be:
65 meters x 4 = 260 x 0.9 = 234 Meters
If you had a 195 foot loss at 3500 rpm. the loss at 1750 rpm. would be: 195 x 0.25 = 48.75 / 90% = 54.17 feet of head loss.
The wear rate of the components varies by the cube of the number.
At 1750 rpm. the impeller material is wearing at the rate of 0.020 inches per month. At 3500 rpm the rate would increase to:
0.020 " x 8 or 0.160 inches per month. Likewise a decrease in speed would decrease the wear rate eight times as much.
I started this discussion by stating that a change in impeller speed or a change in impeller diameter has approximately the same affect. This is true only if you decrease the impeller diameter to a maximum of 10%. As you cut down the impeller diameter, the housing is not coming down in size so the affinity laws do not remain accurate below this 10% maximum number.
The affinity laws remain accurate for speed changes and this is important to remember when we convert from stuffing box packing to a balanced mechanical seal. After the conversion to a mechanical seal we sometimes experience an increase in motor speed rather than a drop in amperage. The affinity laws will help you to predict the final outcome of the change.
The affinity laws also explain the effect on capacity and head when you change motor speed with a variable frequency motor (variable speed driver).
You can use the following formulas to supplement the Affinity Laws. Please keep in mind that these numbers are based on the fluid flowing through the correct size clean pipe.
Product build-up and pipe roughness are variables that will affect the final figures so consider the following "ball park" rather than exact numbers.
Please use these keys when you read the following ratios:
- hf1 The friction loss in the piping, valves and fittings before the change in flow.
- hf2 The friction loss in the piping, valves and fittings after the change in flow.
- Q1 The pump capacity before the change in flow.
- Q2 The pump capacity after the change in flow.
- H1 The pump head before the change.
- H2 The pump head after the change.
- D1 The impeller diameter before the change.
- D2 The impeller diameter after the change.
If you are not familiar with raising a number to some power, please look at the following examples:
32 means 3 x 3 = 9
35 means 3 x 3 x 3 x 3 x 3 = 2433
32.5 means to multiply the square of 3 (9) by the square root of 3 (1.732) = 15.6.
The piping friction loss will vary as the square of the capacity ratio
Example: assume you looked at the friction loss charts and learned that 300 gpm. flowing through a pipeline will suffer 20 feet of friction head loss. Then 500 gpm through the same line will lose:
= 56 feet of head loss.
The pump's capacity varies as the square root of the head on the liquid
Example: if a 160 foot head would deliver 300 gpm. through a specified pipeline, a 100 foot head would deliver:
= 237 gpm
The friction loss in the piping is inversely proportional to the fifth power of the pipe diameter ratio
Example: Assume a 3 inch diameter pipe can handle 300 gpm with a 20 foot friction loss. The same flow rate through a 2 inch diameter pipe would create:
= 152 foot loss
The same flow through a 4 inch line would create:
= 5 foot loss
The capacity of a pipe would vary as the 2.5 power of the diameter ratio
Example: assume that a 3 inch diameter discharge pipe delivers 300 gpm. under a specified head. Under the same head, a 2 inch pipe will deliver:
= 109 gpm.
CO12.1 CARBONIC ACID
This acid forms when carbon dioxide combines with water to form carbonic acid,
CO2 + H2O = H2CO3
The affect is to lower the pH of boiler water causing the operator to increase the amount of water treatment to maintain the proper pH.
The carbon dioxide enters the condensate system through pump packing, the stem of valves above the water line and gasketed flanges.
D035. DYNAMIC BALANCE
Everyone agrees that balancing the rotating components of a centrifugal pump is a good idea, but it is seldom done. Evidently it doesn't appear to be too important or it would be receiving some sort of priority when the pump is being overhauled or rebuilt.
To understand the importance of dynamic balance visualize yourself going down the highway in your automobile at sixty miles an hour, and you throw off a small, lead wheel weight; suddenly you notice a severe vibration in the steering wheel that makes you feel very uncomfortable. Do you have any idea how many rpms the wheels were making at sixty miles an hour or 100 kilometer per hour? Do you think it was slower or faster than the rpm of your centrifugal pump? Let's figure it out in the inch size and then we will do it in metric:
A typical fourteen inch automobile wheel has a tire that is approximately twenty five inches in diameter. This means that the circumference of that tire is 25 inches times 3.14 (pi) or 78.5 inches. Divide the 78.5 inches by 12 and you get 6.5 feet for the circumference of the tire.
At sixty miles an hour you car is going a mile a minute or 5280 feet a minute. Since the 6.5 feet represents one revolution of the wheel we divide that into the 5280 feet and we get 812 rpm at sixty miles an hour.
A common metric tire would have a diameter of 635 mm. Multiply that by 3.14 and you would get just a little bit less than two meters for the circumference. At 100 Km/hr you would be going 1.7 Km or 1700 meters/ minute.
1700 divided by two meters for one revolution of the wheel is 850 rpm.
This means that if a small lead weight can become that significant at 812 rpm or 850 rpm what does an out of balance shaft in your pump do at electric motor speeds?
- The bearings will experience higher loading that will translate to premature failure.
- The mechanical seal faces can separate because of the induced vibration and shaft run out.
- The seal faces can become damaged as the vibration causes the carbon to bounce against the hard face. This bouncing can cause a chipping of the carbon outside diameter.
- Seal drive lugs will experience premature wear.
- Shaft fretting will increase dramatically under bearing grease seals and original equipment mechanical seal designs.
What cause a pump shaft to go out of dynamic balance assuming it was balanced at pump assembly?
- Product attaches to the impeller.
- Impeller wear and damage caused by erosion and corrosion.
- A seal or sleeve that is not concentric to the shaft. Set-screws guarantee that the seal will not be concentric to the shaft or sleeve.
- The coupling, impeller, mechanical seal, bearings, sleeve, keys, etc. were not balanced as an assembly.
- The impeller diameter was reduced and not rebalanced. Impeller material is not homogeneous.
Balancing the impeller
- Be sure to balance the impeller after repair, or coating,
- A two-plane balance is preferred.
- Balance impellers to Quality grade G 6.3 based on the ISI 1040 and ANSI S2.19 &endash; 1975 specification.
- Use quality grade G 2.5 for turbine driven pumps and large armatures. Precision bearing replacement
- On a multi stage pump, balance each impeller separately and then balance them as an assembly
You can purchase two plane balancing equipment to do in house dynamic balancing or you can send the assembly to an outside vendor. The newer electronic vibration analysis equipment has a balancing program available for an additional charge over the vibration package. Check with your vendor for its availability.If you send the rotating assembly to an outside vendor for balancing be aware of several potential problems:
- Unlike automobile tire balancing it is not simple to calculate how much weight to remove and exactly where to remove it.
- Many of these people are experienced in only balancing electric motor armatures where it is common to remove weight by removing portions of the fan blades.
- Outside vendors frequently do not know how to remove weight from an impeller. I have seen instances where weld bead was added to the impeller, or holes drilled through the shroud that interfered with the pump hydraulics.
A small low head, axial flow impeller that attaches to the conventional impeller of a centrifugal pump to increase the pump's suction head (pressure), and prevent cavitation problems.
The main purpose of the inducer is not to generate an appreciable portion of the total pump head, but to increase the suction pressure of a conventional impeller. They reduce NPSH required and permit the pump to run at a higher speed, with a given NPSH available.
An inducer should be specified any time the calculated suction specific speed number is above 18,000 in USCS units or 11,000 in SI units. The inducer will reduce the net positive suction head required (NPSHR) of the pump or permit the pump to run at a higher speed.
The inducer flow angle is some where between five and ten degrees with typically two vanes and no more than four. Inducers have been used successfully with suction specific speed numbers of approximately 24,000 (14,700 metric).
Although the efficiency of the inducer is low, it will not reduce the pump overall efficiency significantly.
Not all pump manufacturers have this feature available
Magnetite is the common names for black ferric oxide (Fe3O4 ), a protective coating the forms on the inside of iron pipe to reduce further rapid corrosion.
Magnetite is sold commercially as a polishing compound and is known as "black rouge".
Once the magnetite forms on the mechanical seal sliding parts several events can occur that will fail the mechanical seal prematurely:
- The sliding parts will experience premature wear.
- The hard oxide film will restrict movement of the sliding components causing the lapped seal faces to open and the hard particles to penetrate between the faces and stick into the softer carbon face.
- Magnetite will stick into the sliding elastomer (O-ring) causing further damage to the close tolerance-sliding surface.
A red form of magnetite, hematite, (Fe2O3) is also found on the inside of the iron piping we often find in hot water systems. Like the black version (Fe3O4 ) it is also very abrasive.
M029. MOTOR SELECTION
Electric motors operate at their best power factor and efficiency when fully loaded so you do not want to purchase a motor that is too big, and common sense dictates that one that is too small is even worse.
In the following paragraphs we are going to learn how to select the correct motor for your centrifugal pump application.
Let's assume we will be selecting the motor for the pump described by the pump curve shown below.
The first thing we must do is decide what diameter impeller we will be using. The above curve shows impeller diameters from "A " to "E".
I have selected letters rather than numbers so that we can work the examples in either metric or inch units.
For our example we will use impeller size "A". You will want to look to the right hand side of the curve to select the last efficiency line. In this case it is the 50% line. This will give you the maximum capacity for that size impeller.
Note the capacity at this point (400) and then transfer this capacity and impeller size to a second graph (many times this information is part of the pump curve or located very close to the pump curve) that is supplied by the pump manufacturer.
The second graph will look something like the one illustrated below.
The numbers on the left side of the graph show either the brake horsepower or the kilowatts being consumed. You can select the appropriate units for your application. According to this graph we will be using about 20 (brake horsepower or kilowatts) at the last efficiency line (400).
One assumption we made during this selection process was that the specific gravity of the fluid we were pumping was one (like fresh water). If the fluid has a higher or lower specific gravity we must multiply the number on the left-hand side of the graph by the specific gravity of the fluid to get the correct horsepower or kilowatts for your application.
If the pump were sized correctly for the application it would run within ten percent of its best efficiency point. For impeller size "A" that would be approximately 325 (as shown on the first graph) so we are going to take advantage of the pump service factor (I'll explain that in a few minutes) to give us the needed horse power if we should occasionally run at this higher capacity (400) or get into any other temporary overload condition such as starting a pump that is rotating backwards.
The service factor rating is supplied by the motor manufacturer and is usually available in three ranges:
- A service factor of: 1.00 / 1.10 - most of these are older motors and a majority of them have undesirable aluminum windings.
- A service factor of 1.15 - this is the most common service factor used in modern motors.
- A service factor of 2.00/ 2.50 - These motors are seldom in stock and have to be built at a premium price.
Motors are available in a variety of horsepower and kilowatt ratings. Typical horsepower ratings would be: 0.5, 1.0, 1.5, 2.0, 3.0, 5.0, 7.0, 10, 15, 20, 30, etc.
Our graph showed that we needed a 17 horse power motor, but a 15 horse power motor will work in this application because of the service factor (15 x 1.15 = 17.25 horsepower available). Keep in mind that any heat generation computations made by the motor manufacturer were made for the motor when it was running at its rated horsepower and not at the service factor rating. All this means is that the motor will run hotter than anticipated, but still within acceptable limits.
Oil refinery applications use a second factor recommended by the American Petroleum Institute (API.). This organization specifies that the factor should be used as an additional safety margin. These factors are:
- To 25 horsepower (18,7 kw.) = 1.25
- From 30 to 70 horsepower (22,4 to 52,2 kw.) = 1.15
- A 100 horse power (74.6 kw.) or more = 1.10
If we take the same example as noted above, and insert the API (American Petroleum Institute) additional requirement, we would come up with:
- If 20 horsepower is needed x 1.25 (A.P.I. specification) = 25 horsepower needed.
There are instances where you can combine the two service factors and come up with a compromise. As an example, suppose that the horse power requirement was 8.7 instead of the 20:
According to the A.P.I. (American Petroleum Institute) you would need 8.7 x 1.25 = 10.8 horsepower, so you would have to go to a 15 horse power motor because there is nothing in between 10 and 15 horsepower. According to the above information a 10 horse power motor has a service factor rating of 1.15 so, 10 x 1.15 = 11.5 horsepower or more than enough to satisfy the API (American Petroleum Institute) recommendation.
Electric motors are sized considering the specific gravity of the liquid being pumped. If a low specific gravity pump is tested with water, or any higher specific gravity fluid, the increase in motor amperage could burn out the motor.
N002. NPSHR, CENTRIFUGAL PUMPS
Net positive suction head required (NPSHR). This is the minimum head required to stop the pump from caviating when it is pumping cold water.
The pump curve that came with your pump shows the NPSH required for any given impeller size and capacity.
This number was determined by pumping cold water through the pump while reducing the suction head, until the pump showed a reduction in discharge head of three percent (3%) due to the low suction head and any formation of bubbles within the pump.
This point is called "the point of incipient cavitation".
ROTARY PUMPS NPIP (Net Positive Inlet Pressure)
Positive displacement or rotary pump people do not use the term "head", they use the term "pressure" instead so NPSHR would not be an appropriate term.
Rotary pumps are often selected to move liquids with a low vapor pressure point, or fluids with a lot of entrained bubbles.
This means that NPIP required (NPSH in centrifugal pumps) is difficult to test. The Hydraulic Institute establishes the point at the first indication of any of the following.
- Cavitation noise is heard.
- A 5% reduction in capacity at constant differential pressure and speed
- A 5% reduction in power consumption at constant differential pressure and speed.
O003_1. APPROXIMATE FLOW THROUGH AN ORIFICE
There are several reasons you might want to install a restrictive device or orifice in a piping system.
- To create a false head for a centrifugal pump, allowing you to run the pump close to its BEP.
- To increase the line pressure.
- To decrease the flow through a line.
- To increase the fluid velocity in a line.
The equation for flow through an orifice is a simple one to understand. Only the units are somewhat awkward.
Q = AV
- Q = The flow in cubic feet per second (ft3/sec).
- A = The area of the orifice in square feet (ft2).
- V = The velocity of the liquid in feet per second (ft/sec).
Experience shows that the actual flow is quite different than calculated because of the different shapes of the various orifices. Look at the diagram on the following page and you will see some of these popular shapes. Each has been assigned a "K" value.
We will enter that "K" value into our equation and the new equation becomes:
Q = AVK
To make the equation easier to handle we can express the velocity "V" as:
- g = 32.2 ft/sec2
- h = Head across the orifice. If the downstream side of the orifice is pressurized use the differential head across the orifice.
If you do not know how to convert pressue to head, use this formula:
It would also make sense to convert some of the terms in our equation to terms that are more convenient to use. As an example:
- "Q" can be converted from cubic feet per second to gallons per minute:
- ft3/sec = 448.8 gpm.
- "A" The area in square feet can be converted to square inches:
- ft2 = 144 inches2
Putting all of this together gives us a new formula that looks like this:
Let's plug in some numbers and calculate a flow through a typical orifice.
- h = 20 feet
- A = 0.049 square inches
- K = 0.62
Q = 25 x 0.049 x 0.62 x 4.47
Q = 3.40 gallons per minute
If we want to solve for the orifice area:
If you are uncomfortable working with the orifice area in square inches you can use the diameter instead. Use the following equation:
Inserting the 0.049 square inches we calculated from the prior formula we get
We made our formula more user friendly by substituting some conversions and now we can make our calculations in gallons per minute and square inches, but the formulas would be better if we could measure the orifice diameter rather than the orifice area
I took you through this exercise to show you how the formulas we use in these papers are derived. We will re-write the flow and orifice diameter formulas again and maybe this time they will be simple enough for anybody to use. We will start with the flow formula and then fix the orifice formula:
The formula for calculating the orifice diameter becomes:
Let's see if the formulas still work. Here are the numbers:
- d = .250 or 1/4inch
- K = 0.620
- Q = 3.4 gallons per minute
- h = 20 feet
We will begin by solving for flow (Q)
Well that worked, now let's try for orifice size:
All of these above numbers were generated assuming that you were moving water through the orifice. If you are making calculations for a liquid other than water you will have to factor in the viscosity of that liquid compared to water.
We also made an assumption that the orifice diameter is not greater than 30% of the pipe diameter. There is another formula we use for a less restrictive orifice.
Any time the ratio of the orifice diameter to the pipe diameter is greater than 30% (0.30) you should modify the formula. The modifier (M) looks like this:
- d1 = orifice diameter
- d2 = pipe diameter
When you are using the modifier, the formulas look like this:
Now we will see what happens when a 0.250 inch (1/4) orifice is put into a smaller cross section 0.500 inch (1/2) pipe, assuming the other numbers stay the same:
This means that you would have to multiply by 1.03, so the 3.46 gpm we got in the last calculation would become 3.56 gpm.
How accurate are these predicted numbers? Anytime you make a calculation using flow as a as part of the equation, you will run into some variables that will affect your results:
- The roughness of the piping walls affects the friction loses.
- The piping material and allowable wall thickness tolerances.
- Solids buildup inside the piping. Calcium in water applications and coke in hot oil applications are typical. Higher temperature usually hastens the solids buildup.
P015. PIPING RECOMMENDATIONS
There are entire books written on this subject so in the next few paragraphs I will try to give you some of the highlights of what you should know about piping as it relates to pumps and seals.
- 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.
- 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.
- Piping should be arranged with as few bends as possible. If bends are necessary use a long radius when ever possible
- Substituting a globe valve for a gate valve in a piping system is similar to adding almost 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 best efficiency point (BEP) with a resultant shaft bending. On the suction side of the pump it will probably cause cavitation problems.
- 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 side down.
- 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 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.
- 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.
- Suction piping should be at least one size larger than the suction flange at the pump.
- 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.
- A check valve and a stop valve should be installed in the discharge line with the check valve placed between the pump and the stop valve to protect the pump from reverse flow and excessive back pressure.
- Expansion joints should be installed between the check valve and the pump.
- Suction and discharge piping should be supported to prevent transmitting forces and bending moments o the pump casing.
- 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 one 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 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. Vortexing can occur if any of the following conditions are present:
- Low liquid levels
- Liquid level falling greater than 3 ft./sec. (1 meter/ sec.)
- There is a large concentration of dissolved gases in the liquid.
- High outlet velocities in pipes leaving vessels. Generally greater than 10 feet/sec. (3 meters/sec.)
- Liquids near their vapor point.
- High circulation caused by asymmetrical inlet or outlet conditions.
- Inlet piping too close to the wall or bottom of the tank. Consult the Hydraulic Institute Manual or a similar publication for recommended clearances.
- In a mixer, the liquid level must be at least one and one half diameters of the blade, above the blade.
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)
- 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).
- 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)
- Velocities of no more than 10 feet (3 meters) per second are recommended in the suction side piping to prevent abrasive wear.
In multiple pump arrangements we would prefer to have the suction bells in separate bays so that one pump suction will not interfere with another. If this is not practical, a number of units can be installed in a single large sump provided that:
- The pumps are located in a line perpendicular to the approaching flow.
- There must be a minimum spacing of at least two suction diameters between pump centerlines.
- All pumps are running.
- The upstream conditions should have a minimum straight run of ten pipe diameters to provide uniform flow to the suction bells.
- Each pump capacity must be less than 15,000 gpm.
- Back wall clearance distance to the centerline of the pump must be at least 0.75 of the suction diameter.
- Bottom clearance should be approximately 0.30 of the suction diameter
- The minimum submergence should be as follows:
- 20,000 gpm 4 feet
- 100,000 gpm 8 feet
- 180,000 gpm 10 feet
- 200,000 gpm 11 feet
- 250,000 gpm 12 feet
The metric numbers are:
- 4,500 m3/hr = 1.2 meters
- 22,500 m3/hr = 2.5 meters
- 40,000 m3/hr = 3.0 meters
- 45000 m3/hr = 3.4 meters
- 55,000m3/hr = 3.7 meters
S072. SPECIFIC SPEED
Specific speed is a term used to describe the geometry (shape) of a pump impeller. People responsible for the selection of the correct size pump 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 net positive suction head required (NPSHR) numbers.
- 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, efficiency 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 published pump curves at the pump's best efficiency point (BEP):
- NS = Specific speed
- N = Pump shaft speed
- Q = Capacity in GPM. For a double suction pump use one half the capacity.
- H = Total head, developed by the largest impeller, in feet
The following chart gives you a graphic picture of the impeller shape represented by this number:
The major use of the specific speed number is to help you specify pumps that are more efficient.
- The maximum pump efficiency is obtained in the specific speed range of 2000 to 3000.
- Pumps for high head low capacity occupy the range 500 to 1000. While low head, high capacity pumps may have a specific speed of 15,000 or larger.
- For a given head and capacity the good news is that the pump having the highest specific speed, that will meet the requirements, probably will be the smallest size and the least expensive. The bad news it that the pump will run at the highest speed where abrasive wear and cavitation damage become a problem.
- Efficiencies start dropping drastically at specific speeds below 1000. Also smaller capacities exhibit lower efficiencies than higher capacities at all specific speeds.
- In propeller and other high specific speed impellers (axial flow) it is not practical to use a volute casing. Instead, the impeller is enclosed in a pipe like casing.
- The lower the specific speed number, the higher the power loss you get with wear ring clearance.
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 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 = gpm, and H = feet, divide the NS by 1.63
British ...................Q = Imp. Gpm, 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 gpm.
- 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 net positive suction head required (NPSHR).
- 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.
The following diagram illustrates the relationship between specific speed (Ns) 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 best efficiency point (BEP) requirement.
- 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.
Here is another curve to show you the relationship between specific speed, capacity and horsepower requirements:
Keep in mind that efficiency and power consumption were calculated at the best efficiency point (BEP). 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 best efficiency point, 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 clearance between the impeller and the tongue of the volute has a bearing on efficiency, pressure pulsations and cavitation. For high efficiency you would want a small clearance, but this produces larger pressure pulsations and the increased flow in this area can reduce the fluid pressure enough to cause flashing of the product and a type of cavitation known as The vane passing syndrome.
For impellers up to fourteen inches in diameter (355 mm) this clearance should be a minimum of four percent of the impeller diameter. If you are using greater than fourteen-inch diameter impellers the clearance should be at least six percent of the impeller diameter. Also remember that as this clearance increases the impeller experiences some slippage. That is the major reason that we do not like to remove more than ten percent of the impeller diameter when trimming is called for.
If you work in both metric and imperial units, as I do, the subject of specific speed becomes very confusing because both systems use the same specific speed numbers to describe the impeller shape. They do this even though they use a different set of units to arrive at the same number.
SA005_1. TEMPERATURE LIMITS FOR COMMON ELASTOMERS
F. TEMPT. RANGE
C. TEMPT. RANGE
V026. VORTEXING LIQUID
Vortexing of the fluid in a suction sump or pit sounds a lot like cavitation problems and will cause excessive shaft deflection that is harmful to:
- Mechanical seals
- The pump intake structure and piping.
One way to tell if you have a cavitation or vortexing problem is to remember that vortexing problems are intermittent as the vortices form. Cavitation once started tends to stay with you. Proper pit or sump design can eliminate this vortexing problem, but what do you do if the installation is not new and the problem exists? There could be several things that could have caused the vortexing problem:
- The pump capacity has increased
- If the discharge head of a centrifugal pump is reduced the capacity will increase.
- Maybe a larger pump has replaced a smaller pump that was originally installed.
- The pump could be running at a faster speed than original design.
- Additional pumps have been installed in the pit.
- The flow or volume to the pump inlet has changed.
- The fluids-solids mixer has changed.
- The pit inlet has been reduced.
- The line is restricted with solids of some type
- You have more air in the liquid.
- The return line is giving a water fall affect.
- A clogged trash rack or screen can restrict some of the incoming liquid.
Maybe the original design was bad and that is causing the problem. Although this is a very large subject there are a few guide lines you might check out:
- To prevent vortexing, the minimum submergence for a continuous running pump is 1.75 times the diameter of the bell (not the pump) inlet. This can vary with pump manufacturers because there is also the possibility of cavitating if you do not have enough NPSH available.
- The pump suction bell should be a minimum of 0.5 diameters off the sump or pit floor.
- The pit inlet should be as far away from the pump suction as possible.
- The usable pit volume should equal or exceed the maximum capacity to be pumped in two minutes.
- If the pumps are on a float switch they should be sized to allow no more than four starts per hour per pump.
The hydraulic Institute ANSI recommendation: ANSI/HI 9.8-1998 Section 9.8.7 recommends the following formula to determine the minimum submergence of a submersible pump::
- S = Minimum submergence to prevent vortexing, in inches
- D = Pump suction inlet diameter, in inches
- Q = Pump design flow rate, in USGPM
Now we will take a look at what you can do with an existing installation. Remember that a low velocity and straight line flow to all pumps is always desired. If you are getting vortexing problems you might be able to:
- Place a cone under the bell.
- Use diffuser screens.
- Use floating rafts around the pump column to break up the vortices.
- Float large spheres on the surface to break up vortices.
- Move the pump away from the wall.
- Reduce the inlet velocity by spreading the flow over a larger area, or change the direction and velocity of the flow by the use of baffles.
- Eliminate the separating wall between pumps.
- Keep the inlet flow to the pit below 2 feet/second (0.7 meters/sec)
- Keep the flow in the pit below 1 foot/sec (0.3 meters/sec)
- Any type of a logical flow straightener will help reduce velocity.
In the next few illustrations I will show you the recommended sump dimensions to prevent vortexing and eddy flows.
The first chart shows the recommended dimensions:
The next two charts show where the dimensions came from:
- Dimensions "Y and A" are recommended minimum values. They can be as large as desired but should be limited to the restrictions shown on the chart.
- If the design does not include a screen, or if the channel has a sloping approach, dimension "A" should be up to two times as long.
- If the channel approach has a down slope the angle should not be more than 15 degrees
About the screens:
- The screen or gate width should not be less than "S". Heights should not be less than "H".
- Use dimension "S" for the width of an individual pump cell, or the center to center distance of two pumps if no division walls are present.
W007. WEAR RING CLEARANCE
Wear rings should be replaced when their clearance doubles. This additional clearance will increase the pump power requirements with the amount of power varying according to the specific speed (NS ) of the impeller
- NS 200 14% increase
- NS 500 7% increase
- NS 2500 Insignificant increase
If the wear-ring clearance is too large the pump will take on excessive vibration caused by internal recirculation. This can cause seal and bearing component damage. Another problem is that the pump will not meet its designed capacity because of the internal recirculation.
When replacing wear rings, it is not uncommon to bore out the stationary ring and machine the rotating ring oversize to get the correct clearance. If you do this on a double-ended pump, be sure to do both sides of the impeller to prevent upsetting the balanced hydraulic forces and thrusting the impeller to the end that was not bored oversize.
Also be aware that many wear rings supplied by OEM manufacturers are out of round and must be machined to get the proper clearance inside the stationary ring
P042 SOME COMMON MISCONCEPTIONS ABOUT CENTRIFUGAL PUMPS
Eliminate the paragraph: You should never throttle the suction of a centrifugal pump
ans. If the product you are pumping is explosive and close to its vapor point, suction throttling my be your only option. Discharge throtteling would produce additional heat that might be dangeous.
F040 The most asked questions about mechanical seals
What is considered good life for a mechanical seal?
- The only part of a mechanical seal that is supposed to be sacrificial is the carbon face. The seal should run leak free until the carbon face is worn away. If the seal leaks for any other reason we consider it a premature failure and always correctable.
- Two hard faces are selected when carbon is not acceptable in the application and you have run out of options. You are then trying to get the longest life you can.
- The only variable in seal life should be the lubricating quality of the product you are sealing. Hot water, many gases and most solvents are typical non-lubricants.
- With all of that said, the fact is that in excess of eighty-five percent of mechanical seals fail prematurely. When seals are removed from the running pump most of the carbon face is still intact. Little face wear is the rule not the exception.
Why do most seals fail prematurely?
- One of the seal components becomes damaged.
- The seal faces open.
What are the most common causes of component damage?
- Corrosion of one of the seal components.
- Physical damage that includes the affects of high heat or excessive pressure
What are the most common causes for the lapped seal faces to open?
- The seal was set screwed to a hardened shaft.
- Solids in the product you are sealing are clogging the moveable components.
- The product changed state and interfered with the free movement of the seal. It:
- Became viscous.
- Built a film on the sliding components and the lapped faces.
- The product vaporized across the lapped faces blowing them open.
Do seal faces have to be lubricated? Can they run dry?
- The graphite in the carbon/graphite face is a natural lubricant. In operation the graphite separates from the mixture and transfers to the hard face. This means that the seal face combination you are normally running is carbon on graphite. The hard face is just some place to put the graphite.
- Moisture must be present for the graphite to separate from the carbon/graphite mixture.
- Running dry means higher heat at the faces. If you are using a good unfilled carbon/graphite (and you should be) the faces are not going to be your problem. The elastomer and the product you are sealing can be very sensitive to a temperature change in the stuffing box, or an increase of temperature at the seal faces.
Do seal faces have to be kept cool?
- Most carbons and hard faces can tolerate a lot of heat. The elastomers (rubber parts) are the parts you have to watch. They are the most sensitive to a change in stuffing box temperature, especially if they are positioned in the seal face.
- Hydraulically balanced seals generate very little heat between the faces.
- Unbalanced seals usually require cooling because of the excessive heat they can generate.
- Some face combinations generate more heat than others. Two hard faces as an example.
- Some seal materials conduct heat better than others. Ceramic is a poor heat conductor and carbon is not much better. Tungsten carbide and silicone carbide are excellent conductors of heat.
When should you use two hard faces?
- With any of the oxidizing agents.
- When sealing any of the halogens.
- If the product tends to stick the faces together.
- If you are sealing hot oil and you have to pass a fugitive emission test.
- Some de-ionized water will attack carbon in any form.
- When you are not allowed anything black in the system because of the possibility of color contamination.
- Any time carbon/graphite will not work for some reason.
- If the specifications call for two hard faces.
Why not standardize on two hard faces?
- They generate higher heat than the carbon/ hard face combination.
- They are not very forgiving. If the faces are not dead flat at installation, they seldom lap them selves flat in operation.
Do seals have to leak.?
- Any good quality mechanical seal should run without visible leakage.
- Single, stationary, (the springs do not rotate) hydraulically balanced mechanical seals can pass a fugitive emission test as long as the rotating portion of the seal is designed to be located square to the shaft.
- Rotating seals (the springs rotate with the shaft) seldom can pass a fugitive emission test. They are too sensitive to various forms of misalignment.
- Cartridge mounted stationary seals usually fail fugitive emission testing because the set screwing of the cartridge to the shaft prevents the rotating face from positioning its self square to the shaft. Some seal companies offer some type of a self aligning design to solve this problem.
Why do most original equipment seal designs frett and damage the shaft under the dynamic elastomer or spring loaded Teflon.?
- Corrosion resistant shafts and sleeves protect themselves from corrosion by forming a protective oxide (ceramic) layer on the metal surface. The dynamic elastomer in the seal polishes this layer away as the shaft slides through the elastomer because of shaft vibration, pipe strain, misalignment etc.
- The ceramic protective oxide that is removed by the polishing action imbeds its self into the elastomer causing it to act as a grinding wheel that increases the sleeve or shaft damage.
Do you have to flush most slurry applications?
- It depends upon the percentage of solids. Most fluid with entrained solids can run without flush if you have met the following conditions:
- The packing stuffing box has been replaced with a larger inside diameter version. Centrifugal force will throw the solids away from the lapped seal faces.
- You are using a hydraulically balanced seal that generates low heat.
- The seal springs are not located in the fluid.
- The fluid is at the seal outside diameter.
- The dynamic elastomer moves to a clean surface as the carbon wears.
- You are using suction recirculation to get flow in the stuffing box.
I am looking for a simple solution to a difficult problem. Do discharge recirculation filters or cyclone separators installed between the pump discharge and the stuffing box make sense in slurry applications?
- I wish they did!
- Filters clog and then there is no circulation in the stuffing box.
- Cyclone separators were never intended to be a single pass devise. The also require a substantial difference in pressure between the discharge and the clean liquid connections. In a pump application these pressures are too close together.
If I put a higher fluid pressure barrier fluid between dual seals, shouldn't that keep the faces clean?
- No, the clean fluid always takes the path of least resistance. That is the same reason that higher pressure air does not keep dry solids from penetrating the lapped faces.
- Centrifugal force will pack solids in front of the inboard seal face and restrict its movement.
Do you need a higher pressure barrier fluid between dual seals?
- Higher pressure is called barrier fluid, lower pressure is called buffer fluid.
- The only dual seals that require a barrier fluid are the "back to back" rotating, unbalanced versions, and you shouldn't use them any way.
- Balanced tandem seals (one behind the other) use a buffer fluid that will not dilute your product if the inner seal fails. They also put the pumping fluid at the inner seal outside diameter where it belongs.
- Dual seals should be hydraulically balanced in both directions so that they will stay shut regardless of the direction of the fluid pressure.
How does seal hydraulic balance work?
- There are two forces closing the seal faces.
- A spring force caused by the spring, springs, or bellows pushing on the seal face.
- A hydraulic force caused by the pressure of the fluid acting on the closing area of the seal faces.
- There are three forces opening the mechanical seal:
- A hydraulic force caused by fluid or vapor trapped between the lapped faces.
- Centrifugal force that is causing the rotating portion of the seal to try and become perpendicular to the rotating shaft.
- Hydrodynamic forces generated between the seal faces because for all practical purposes liquids are not compressible.
- We balance these forces by reducing the closing area of the seal faces and thereby reduce the closing force. This is usually done by a small sleeve inserted into the seal or as step machined into the shaft. Metal bellows seals have an effective diameter measured through the bellows to accomplish the same thing.
Is it O.K. to have a third party rebuild my mechanical seals?
- Not really. If you're happy with your seal have the manufacturer, or the company that sold it to you do the rebuilding. Here are a couple of reasons why:
- Carbon/graphite has to be molded in a sintering process and the third party doesn't own the molds for your carbon/graphite face. Machined carbons don't have the density required for good seal faces.
- There are many grades of elastomers. How do you insure you have the right grade. You can't tell by looking at the part.
- Lapping is a real art. The temperature has to be closely controlled to get the right flatness.
Should I be using split mechanical seals?
- There are places where they are the only logical solution:
- Double ended pumps. If one seal is leaking why take the pump apart and change both? Change only the one that is leaking.
- Large vertical pumps. Sometimes you have to take the roof off the building to remove the solid mechanical seal.
- Large size shafts are a natural for split seals.
- Changing a seal means doing a re-alignment. Why go through that again?
- If you have to remove a lot of pump insulation to get to the seal.
- If the pump is in an awkward location, split seals make sense.
- Many split seal designs can run with no visible leakage, but they seldom can pass a fugitive emission test that calls for leak rates in the order of parts per million.
If I touch the lapped faces, are they ruined?
- Not at all. Touching seal faces seldom causes problems. We are trying to keep solids from penetrating between the lapped faces, so the less you handle them the less likely solids will be deposited on the faces.
Why should you not use stainless steel springs or stainless steel bellows in mechanical seals?
- Chloride stress corrosion is the problem and chlorides are every where. Use hastelloy "C" springs and metal bellows and you'll never have this problem.
Why not standardize on Teflon as the preferred rubber part in a mechanical seal?
- Teflon® is not an elastomer, it doesn't have a memory and has to be spring-loaded to the sleeve or shaft. This spring loading interferes with the flexibility of the seal and prevents the elastomer part from flexing and rolling to compensate for minor shaft movements.
Why not mount the seal outside the stuffing box and then dirt and solids will not get into the springs and sliding parts of the mechanical seal?
- The sealing fluid will be at the inside diameter of the lapped faces where centrifugal force will throw solids into the faces.
- Solids will pile up in front of the seal preventing the faces from moving forward when the sacrifical carbon wears.
What is a cartridge seal?
- The rotating portion of the seal is mounted on a cartridge sleeve and this assembly is connected to the stationary portion of the seal along with the seal gland to form a cartridge assembly. Cartridge seals simplify the installation process and allow you to make impeller adjustments without upsetting the seal face loading.
Do I need the new gas seals if I want to seal fugitive emissions?
- Not really. Rotating seals do not pass fugitive emission tests because of their sensitivity to misalignment. Stationary seals usually do not have this limitation.
- The difficulty arises when you try to install a stationary seal on a cartridge sleeve. When you tighten the sleeve set screws to the pump shaft you introduce misalignment between the rotating seal face and the rotating shaft. Hysteresis (delay or lag) problems take over and the result is the stationary seal design fails to pass the fugitive emission test. Any good cartridge mounted self aligning seal can resolve this problem.
- Although a single seal can pass the test, a dual seal is recommended with a low pressure buffer fluid between the seals to act as a back up when the first seal wears out or fails. The buffer fluid will prevent unwanted product dilution and simplify the installation because there is no need for a compatible high pressure barrier fluid that is often hard to find.
Why does my outside mounted seal make a whistling sound?
- The seal faces are running dry. The product you are trying to seal is not a lubricant.
Every time I remove a rubber bellows seal from my pump it is stuck to the shaft. Why?
- It is supposed to vulcanize its self to the shaft so that it can drive the rotating face. If you can remove it easily something is wrong. You probably used the wrong lubricant on the rubber during installation. This is a case where the lubricant we use is supposed to attack the rubber and make it swell.
When my metal bellows seal fails because of breakage at the plates, the break is always near the end fittings and never in the middle of the bellows. How is that explained?
- This is the common mode of failure for excessive vibration. Metal bellows seals need some type of vibration damping to stop harmonic and "slip-stick" vibration problems.
® DuPont Dow elastomer
C075 CORROSION - IRON
Iron, exposed to moist air, will react with oxygen in the air to form iron oxide. This oxidation process is called rusting.
iron + oxygen iron(III) oxide. 4Fe(s) + 3O2(g) 2Fe2O3(s)
Types of corrosion
- Rusting requires both oxygen and water.
- Salt, acid and higher temperatures will accelerate oxidation (rusting.)
- Iron and steel are most commonly protected by painting, plastic coating, and metal plating.
- Galvanizing (Zinc plating) is another alternative
- The oxide formed by oxidation does not firmly adhere to the surface of the metal and flakes off easily causing "pitting." Extensive pitting eventually causes structural weakness and disintegration of the metal.
- Aluminum, forms a very tough oxide coating (Aluminum oxide Al2O3 ) which bonds to the surface of the metal preventing the surface from further exposure to oxygen and corrosion.
- The patina we see on brass and bronze is really corrosion but it protects the metal from further deterioration. In some cases it is considered aesthetically pleasing. The oxidation of copper produces a protective coating of copper oxide, which is red, but then thickens to give a familiar green patina and/or the dark brownish color of bronze statuary.
Combined effect of mechanical factors and corrosion.
General corrosion / uniform corrosion. Metal corrodes uniformly all over the surface.
- Graphitization. Gray iron is a matrix of iron and graphite that acts as a lubricant during machining. A galvanic action can take place between these dissimilar materials that will cause the iron to gradually go into solution, leaving the graphite behind. This can cause the casting to break off into pieces.
- Selective leaching can occur when di-ionized water circulates through pipes and pumps and extracts those minerals that were removed during the di-ionization process.
Local corrosion. Part of the structure corrodes at a considerably higher than average rate. The categories of local corrosion are:
- Pitting corrosion. The corrosion effect is concentrated on localized areas and leads to pitting.
- Crevice corrosion proceeds at locations covered by a corrosion product and other deposits (dirt or trash). Crevice corrosion typically occurs in small cavities, gaps, recession, etc.
- Galvanic corrosion requires two different metals, constituting a corrosion cell. A structure should contain metals that are as close to each other as possible in the galvanic electric series.
- Intergranular corrosion proceeds along the metal grain boundaries.
- Interfacial corrosion occurs at water-air interfaces.
- Cavitation Erosion occurs when passivated metal pieces are removed as the bubbles collapse, leaving the active metal exposed to corrosion. The process is ongoing.
- Selective corrosion occurs when one element in an alloy dissolves faster than the others.
- Exfoliation corrosion attacks the exposed material end grain and can work its way parallel to the metal surface, creating products of greater volume than the original material and causing splitting of material layers, leading to a stratified appearance
- Mechanical wearing as well as static or dynamic stresses often act in combination with corrosion. The main categories for the combined effect of mechanical factors and corrosion are:
- Stress corrosion occurs when a metal in a corrosive environment is exposed to static or residual stress that results in fracture. Shot peening and heat treating are a couple of the remedies to this type of corrosion
- Corrosion fatigue is caused by the combined effect of corrosion and varying state of stress.
- Erosion corrosion is acceleration in the rate of corrosion caused by high velocity of a liquid, or solid impurities carried by a liquid.
- Cavitation corrosion is erosion caused by the combined effect of corrosion and the pressure caused by the breaking of gas bubbles formed in liquid (cavitation) damaging the passivated coating
- Fretting corrosion occurs between two metals rubbing against each other under corrosive conditions. The rubbing removes the passivated layer that had formed on the metal parts. Lip or grease seals are another common cause for the removal of this protective layer.
C076 Corrosion problems with submersible pumps
- Cast iron pumps solves most submersible pumping operations. This normally presents no corrosion problems when pumping liquids such as surface water and domestic sewage. The low oxygen content in raw sewage lowers its corrosion effect to almost nothing.
- Stainless steel is sometimes used as the main material in submersible pumps for two reasons:
- Corrosion resistance in acidic liquids
- Where the purity and color of the liquid being pumped is critical.
- A less expensive and more flexible alternative against salt-water corrosion is to use a protective coating on a conventional cast iron pump. The most widely used coating is epoxy (a polymer material). Normally, zinc anodes are used in conjunction with epoxy coating because of the inevitability of abrasions and scratches occurring in the coating.
- The thermal coefficient of expansion of most plastics is several times that of cast iron, so watch out for separation at elevated temperatures.
- The use of sacrificial zinc anodes significantly extends the life of a coated pump. Between five to ten anodes are implanted at various points around the cast iron structure of the pump.
- An alternative to using sacrificial anodes is to supply the micro current by cable from an external power source. This is known as an "impressed current." In this method, a non-sacrificial anode is suspended in the liquid beside the pump. The anode is attached by cable to the pump and the micro current is introduced into this cable. The impressed current method, however, is complicated, expensive and requires a lot of monitoring to be effective.
- The extra cables required become obstructions in the pump well and often get damaged.
- The implanted (sacrificial) anode method, being simpler and less expensive, is normally preferred.
- Chlorinated rubber is often used to prevent corrosion of cable sheathing for conventional pumps.
- Fluorinated ethylene plastic is an alternative coating material
S116 Surging pump
Some thoughts on pump surging
- Random surging is caused by an air pocket getting loose in the suction piping
- Dissolved gases in the system
- The pump suction is below atmospheric pressure (vacuum) and air is coming into the packing, Stop it by converting to a mechanical seal.
- The fluid is vortexing at the pump suction. Use a vortex breaker
- An air pocket at the suction caused by a reducer that has been installed upside down
- Constant surging is usually a result of a low NPSHA
- Internal recirculation is heating the liquid to its vapor point
- A stuffing box suction recirculation line is being used with fluid at or near its vapor point
- The well is not keeping up with the pump.
- If you have a pump with a flat curve (low specific speed), a small variation in pressure (head) can cause a significant change in flow rate.
- If the suction side of the pump is under vacuum conditions, a liquid at elevated temperature, close to its vapor point is always a problem
- Is there anywhere in the system that vapor can accumulate and cause random surging as it eventually finds its way to the pump?
- You could be seeing a very high frequency "water hammer," except instead of one or two "hammers" that is usually associated with a water hammer, it is appearing as a high frequency vibration or pulsation.
- At any restriction in the pipe, a valve, or a change of pipe area, a certain portion of any pressure wave is reflected, while the balance continues down the pipe.
- When a restriction is located at a point where the pressures are reflected in "tune," or at a particular harmonic being generated by the pump, vibrations can increase in amplitude.
- It is much more commonly known in reciprocating pumps and compressors, Centrifugal pump pulsation is less commonly known, apparently because it is much less of a problem for most users.
- One method used to avoid flow surging problems is proper location of the discharge throttle valve.
- When a centrifugal pump is operated at a very low flow rate, recirculation occurs within the impeller, and it surges at the natural frequency of the system. As the control valve is moved away from the pump, there is a decrease in the frequency and an increase in the amplitude of the pressure waves
- The energy imparted to the system by the pump is similar to the strumming of a guitar string. The frequency is a function of the length of the string, and is analogous to the distance from the pump to its control valve. The greater the distance between the valve and pump, the lower the frequency of this oscillation and the greater the magnitude of the pressure pulsations.
- Placement of the valve close to the pump discharge flange minimizes the amplitude, and thus the effects of the flow oscillations. The mass of the oscillating fluid is reduced in volume, and the turbulent flow through the valve destroys the frequency of the excitation force.
- If the throttle valve is remotely located, flow surging will be of low frequency and high amplitude since the fluid mass is large. This situation must be avoided, as it may induce violent mechanical vibrations in the piping system.
- The preferred throttle valve location will always be close to the pump discharge flange in order to minimize the potential and the effects of flow surging.
- Flow surging problems can also be resolved by installing a bypass line. Bypassing a portion of the pump capacity back to suction maintains pump operation closer to its design flow where the amplitude of the hydraulic excitation forces is small. The bypass line also protects the pump from overheating and damage if system flow rate is reduced below minimum flow.
- The stiffness of the pumping system affects both the frequency and amplitude of pump surging. A "soft" system can result in low frequency high amplitude surging that is detrimental to pump performance and life.
- A "soft" system can be caused by a number of factors:
- There can be air or gas entrained in the pumping fluid, air can be trapped in an elbow, or a system can have open tanks or an accumulator.
- A closed loop boiler with a remotely located control valve that takes only a minimal pressure drop at low flow rates, or a system in which there were flexible lines attached to the suction and discharge of the pump, would be typical of a soft system.
- The following recommendations should be applied to minimize the effects of soft systems:
- Place the control valve at or within five feet of the discharge of the pump.
- Use flexible discharge lines only downstream of the control valve and use a minimum of 10 pipe diameters of straight piping into the pump suction.
- Size pump control valves so that 5-8% of the differential pump head is taken across the valve.
- Install a low flow bypass system if operation below minimum flow is anticipated
- Eliminate entrained air and gasses, or anticipate a higher minimum flow limit if they cannot be avoided.
- Avoid air traps in pumping systems.
- Consult your pump supplier if operation is anticipated below 40% of the BEP for any sustained period and the NPSH is less than 30 feet.
- In hydraulically soft processes where a capacity control valve cannot be located near the pump, and maximum turndown is needed, a discharge orifice is often recommended. This isolates the pump from potential system resonances, and permits stable operation at flow rates down to 15-20% of BEP (the pump's best efficiency point). The orifice is normally sized to drop 5% of the pump head at BEP.
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