Article #1: How Does Pump Suction Limit the Flow?
One of the claimed advantages of the centrifugal pumps over positive displacement pumps is their ability to operate over a wide range of flow. Since a centrifugal pump operates at the intersection of a pump curve and a system curve, by varying the system curve the operating point of the pump is easily changed:
Figure 1-1 Flow control of the centrifugal pump by the discharge valve
The convenience and simplicity of such flow control by the discharge valve throttling comes at a price, because a pump is thus forced to run either to the left, or to the right, of it's best efficiency point (BEP). However, the real danger of operating the pump too far off-peak comes from the suction side considerations. Too far to the right - and you are easily risking to run out of the available NPSHA, causing cavitation problems. Too far to the left - flow recirculation at the impeller eye will let itself known through the noise, vibration, and damage. Thus, the flow must be limited on both sides of the BEP:
Figure 1-2 Pump operating range has limits
Consider the first limitation - high flow. Centrifugal pump stops pumping when liquid turns to vapor. This happens when the pressure somewhere inside the pump drops below liquid vapor pressure. Vapor pressure depends on the temperature, and a few other things. As we know, water turns to vapor at 212 oF at atmospheric pressure, when we boil water in the open pot. If the pot were closed, the water would reach higher pressure before it boils. Conversely, if the pressure were reduced (vacuum), water would boil at lower temperature. It will boil at room temperature, if the absolute pressure is less then about 0.4 psia. Water has low vapor pressure, but other substances may have very high value.
Freon, for example, has vapor pressure of about 90 psia, and ethane value of vapor pressure is about 700 psi, - at 80 0F. Knowing vapor temperature without relating it to a corresponding temperature is meaningless. Sometimes it is good to have a tabulation, or a graph, showing the relationship between the vapor pressure and temperature. The higher the temperature - the higher the vapor pressure is.
Centrifugal pump is a "pressure generator", produced by the centrifugal force of its rotation impeller. The pressure gets higher as flow progresses from the suction to discharge. This is why vaporization of liquid is most likely to happen in the inlet (suction) region, where the pressure is lowest. In practice, it is difficult to know exactly when vaporization (cavitation) happens, so it is wise to keep some margin of available pressure over vapor pressure. Pressure is expressed in "psi", but can also be expressed in feet of water, and the conversion formula is:
FT = PSI x 2.31 / SG, where SG is specific gravity.
This pressure, expressed in feet of water, is called discharge head at the pump exit side, or suction head on the inlet side. The difference is a pump developed head, also called a total dynamic head (TDH). These heads must include both static and dynamic components. Static part is what we measure by the gage in front of a pump, and dynamic, according to Bernoulli, is velocity head V2/2g.
For example, suppose an inlet pressure gage installed in a 2" pipe directly in front of a pump delivering 100 gpm oil with specific gravity SG = 0.9, reads 10 psig. To calculate velocity head, find the pipe net area, which is A = 3.14 x d2 / 4 = 3.14 x 22 / 4 = 3.1 in2.
The velocity can be calculated by the formula:
V = (Q x 0.321) / A = (100 x 0.321) / 3.1 = 10.4 ft / sec
Then, the velocity head is:
V2 / 2g = 10.42 / (2 x 32.2) = 1.7 ft, or, converted to psi is
= 1.7 x 0.9 / 2.31 = 0.7 psi
The total suction pressure is then 10 + 0.7 = 10.7 psi, or, if expressed in feet of water,
= 10.7 x 2.31 / 0.9 = 27.5 feet
It is best to have gages as close as possible to the pump, on the suction and discharge sides. Unfortunately, often these gages are not installed, (which somehow happens more often on the suction side), and suction head in front of the pump is estimated by calculations, by subtracting the pressure (head) losses from the known value of head upstream, and adjusting by elevation correction, according to Bernoulli. In many cases, the upstream datum is a known liquid level in a suction tank.
a) Tank open to atmosphere:
Figure 1-3a: Open tank
Figure 1-3b: Pressurized tank
hsuction = 2.7x2.31/0.9 + 10 – 7 = 9.9’
Figure 1-3c: Tank under vacuum
For water and similarly low viscosity liquids, suction losses are usually low, and often are disregarded. However, for more viscous substances, such as oils, these losses can be substantial, and may cause the pressure in front of the pump drop below the vapor pressure, causing cavitation. This is why the inlet velocity must be minimized, as the losses depend on velocity squared.
Longer pipe runs, bends, turns and other restrictions, add to inlet losses, leading to further pressure reduction in front of a pump. As a quiz, using the examples above, see if you can figure out what happens to inlet pressure if the pipe diameter is doubled? Or made half the diameter? (If you do – send the answer to us, and will publish it the Pump Magazine).
To avoid cavitation, what matters is not the suction pressure, but how much higher it is then the vapor pressure of the liquid being pumped. This is where a concept of NPSH comes handy. The available NPSHA thus is simply the difference between this total suction head, as discussed above, and vapor pressure, expressed as head, in feet.
Pump manufacturers conduct tests by gradually lowering suction pressure, and observing when things begin to get out of hands. For a while, as pressure decreases (i.e. NPSHA gets smaller), nothing happens, at least nothing obvious. A pump, operating at a set flow, keeps on pumping, and develops constant head. At some point, when the value of suction pressure (and corresponding NPSHA), reaches a certain value, a pump head begins to drop, which typically happens rather suddenly:
Figure 1-4: Development of Cavitation
Actually, things are happening inside the pump well before the sudden drop of head, but they are not as obvious. First, at still substantial suction pressure, small bubbles begin to form. This is called incipient cavitation - sort of tiny bubbles in your water cattle that begins to percolate before water is fully boiling. These small bubbles are formed and collapse, at very high frequency, and can only be detected by the special instrumentation. As pressure is decreased further, more bubbles are formed, and eventually there are so many of them, that the pump inlet becomes "vapor-locked", so that no fluid goes through, and the pump stops pumping - the head drops and disappears quickly. It would be nice if enough pressure was always available at the suction so that no bubbles were formed whatsoever. However, this is not practical, and some compromise must be reached. The Hydraulic Institute (HI) has established a special significance to a particular value of NPSHA, at which the pump total developed head drops by 3%. The value of this NPSHA, at which a pump losses 3% TDH, over (i.e. in access of) vapor pressure is called net positive suction head required (NPSHr) in order to maintain 3% TDH loss.
NPSHr = (Hsuction - Hvapor), required to maintain 3% TDH loss
NPSHr is, therefore, established by actual test, and may vary from one pump design to another.
In contrast, the available NPSHa, has nothing to do with a pump, but is strictly a calculated value of total suction head over vapor pressure. Clearly, NPSHA must be greater then NPSHR, in order for a pump to make its performance, i.e. to deliver a TDH, at a given flow.
It is easy to know when a NPSH problem is obvious - a pump just stops pumping, but the vapor bubbles do not need to be so dramatically developed to cause TDH drop, - even smaller bubbles can cause problems. The evolved bubbles get carried on through the impeller passage, at which pressure is rising from inlet to exit of the blade cascade. This increased pressure causes the reverse to what happened to a bubble "awhile back", when it first became a bubble formed from a liquid particle during phase transformation (boiling). Now, the bubble is at the somewhat higher pressure, which tries to squeeze it, against the vapor surface tension that keeps the bubble a bubble. The bubble collapses (implodes), with a sudden in-rush of surrounding liquid into a vacuum space previously occupied by the bubble. The inrush is accompanied by a tremendous, but a very localized, pressure shock, which, if imploded in the vicinity of the metal (impeller blade), would cause a microscopic hammer-like impact, eroding a small particle of metal. With enough bubbles and enough time, the impeller vanes can be eroded away quickly, a phenomenon known as cavitation (hence the word) damage.
This is why an NPSHA margin (M=NPSHA-NPSHR) is important, which is typically at least 3-5 feet, and preferably should be even more, if possible.
The NPSHR, discussed above, was so far limited to a particular flow on a pump performance curve. At higher flow, the internal fluid velocities are higher, and, according to Bernoulli, the static pressure (or static head) part becomes less, i.e. closer to vapor pressure. The static pressure, therefore, must be increased externally, i.e. a higher value of NPSHR is needed for higher flows. This is why the NPSHR curve shape looks like this:
Figure 1-5: Ample margin of NPSHA is important
It is important to specify an ample margin of NPSHA over the pump NPSHR for a complete range of operation, and not just at a single rated flow point. The following example illustrates a common mistake, leading to the NPSH-problem. The pump was procured with the intend to deliver between 350-500 gpm, and the manufacturer quotation indicated 16 feet required NPSHR at 500 gpm. As a process later changed, more flow was required, and the discharge valve was opened to allow pump to deliver more flow, 750 gpm. However, as can be seen from Figure 1-5, at about 700 gpm, the NPSHR exceeded the NPSHA available at the installation, and pump started to experience typical NPSH problems - noise, loss of performance, and impeller cavitation damage.
An instinctive thought to address the issue of cavitation due to flow-run out operation is to "overkill" on a pump size, and therefore always stay to the left of the BEP. In the example above, a larger pump, having same 16 feet NPSHR, but at 750 - 800 gpm, would never run out of the NPSHA. That is true, and, in fact, this is exactly what has been a common practice in the past, where an oversized (and, by the way, more expensive) pump would be specified "to make sure", - just to discover other, just as severe problems.
When a centrifugal pump operates below certain flow point, a phenomenon known as flow recirculation in the impeller eye starts. This depends on several design factors, such as suction specific speed (see in other article of Pump Magazine), but generally recirculation begins below 80-60% flow, and becomes quite sever below 40-20%. At even lower flows, recirculation may become especially severe, and is known as surge - violent, low-frequency sound, accompanied by strong low-frequency vibration of the pump and piping:
Figure 1-6 Problems come up when pump operates at too low flow
In addition to obvious mechanical problems with recirculation, the flow undergoes a complex vortexing motion at the impeller inlet (eye), with localized high velocities of the vortex causing horse-shoe looking cavitation damage, usually on the "blind" side of the blade, as compared to high-flow cavitation. Other problems add oil to the fire - radial thrust, which sky-rockets at low flow, causes deflections of the shaft, leading to seal leaks, bearings life reduction, and even shaft breakage (see other articles of the Pump Magazine on these subjects).
Troubleshooting methods and failure analysis techniques help to pinpoint a cavitation problem with a particular pump. The indications of the high flow cavitation are different from the low flow recirculation damage. Side of the blades, the extend and shape of the cavitation trough, can be helpful in determining the causes of each individual problem.
To learn more about this topic, e-mail your comments to us at:
Below are some comments from a reader:
Dear Dr. Nelik,
How does a change in pump suction configuration affect the system curve? I've read the Article #1 "How Does Pump Suction Limit the Flow" and I've also read reader question #58 that deals with changes to the inlet and outlet piping to a pump. After carefully reading Article #1, I can conclude that suction side considerations do not truly limit flow, but flow must be adequate to prevent cavitation (too high flow - low pressure - cavitation) or flow recirculation at too low flow. To stay out of these problem areas, one must design the system (downstream of the pump) and select a pump so that the operating point flow rate does not cause upstream problems. My question revolves around how the upstream configuration affects the system curve - if at all. There appears to be a degree of independence between the system configuration downstream of the pump and the upstream configuration, since the only influence on where the system curve moves to is dependant on the total static head (difference in elevations in the fluid levels for the suction and discharge, and any additional pressurization if present). Friction in the suction line does not appear to influence the system curve but it does influence the NPSHA. Reader question #58 describes changes to inlet and outlet piping to support a ship in dry-dock. The first step is to determine where the system curve for the new outlet piping intersects the pump curve and you get an operating point. Then for that flow rate, check the NPSHA for the new suction piping configuration against the NPSHR required for the pump. If NPSHA is greater (with some margin) than NPSHR, you should be OK. This leads me to my question. We have two identical generators in two separate machinery rooms on a ship. For simplicity, let's say that the inlet and outlet piping for the Seawater cooling is the same length for both generators - 20 feet of inlet piping and 20 feet of outlet piping. The only difference is that because of space constraints, one generator has 4" inlet piping to the generator skid, and the other has 6" piping to the skid. When the suction line reaches the skid, the piping is again identical for both - a one-foot long 5" line that enters the pump. The size of the outlet piping from the skid on both is the same. The system curve (dependent on the downstream pipe configuration) should be identical for both. The total static head is identical for both. Therefore based on the methods described above, the pump should operate at the exact same point for both. Since the suction line size is different, the NPSHA to the two pumps is different and the pressure at the pump inlet is different (there are more losses through the 4" suction). Provided there is no cavitation in either case, would this influence the system curve, and therefore the operating point, in any way? Based on what I've read, I think the answer in no. That's why I said before there seems to be a degree of independence between the piping upstream and the piping downstream as far as operating point is concerned. When the units were running, both pumps should be operating at the same outlet pressure and flow, but upstream of the pump, the water would obviously be going faster through the 4" suction pipe. This is what I think will happen, and I want to verify.
Lionel A. Sequeira
Mechanical Engineer, Dept of the Navy
You got the concept pretty well, with just a small correction. Both pumps will impart the same amount of energy to the fluid. (And, technically, pump head is exactly that: energy per unit of mass). Thus, both pumps will show the same differential pressure at a given operating flow (assuming, as you said, no cavitation). However, the first pump will start off at a lower inlet pressure, because it will loose more pressure on friction between the supply tank and pump inlet pipe due to smaller (4”) pipe. Thus, the discharge gage will show lower pressure for the first pump, - by the amount of difference in losses between the first pump suction losses and the second pump suction line losses.
Remember, pump head is not discharge pressure expressed in feet of water, but the differential pressure, expressed in feet of water.
Another way of thinking this is that the pump head is the difference between total energies between discharge side (near the pump) and the suction side (also right at the pump), corrected by gage elevation differences (which is usually minor). The discharge head is a sum of pressure head, velocity head and elevation head. Same goes for the suction. Assuming the gages are at the same elevation, this cancels out. The velocities are the same, as in the very immediate proximity (discharge flange at exit, or a 5” short pipe at suction) pipe sizes are the same, thus velocity heads are the same for a given flow. Since the total differential is the same, then the pressure difference is the same as well. But, the suction inlet pressure is lower, and, adding the constant differential amount to it makes the discharge pressure lower for the first pump.
I am impressed that you are
getting really deep into pump hydraulics! As a suggestion – also consider to
Keep on pumping!
Dr. Lev Nelik, P.E.
Editor, Pump Magazine