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.
Examples:
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.
Regards,
Lionel A. Sequeira
Mechanical Engineer, Dept of the Navy
9/30/04
Editor comments:
Lionel:
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
attend our
Keep on pumping!
Dr. Lev Nelik, P.E.
Editor, Pump Magazine