PUMP MAGAZINE: Questions and Answers
(31-50)
Editorial staff
continuously updates Q&A section by adding new questions and answers, based
on our readers’ interest, input and feedback.
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Question 31: Dear Dr Pump,
We have a Progressive Cavity pump and working to
discharge Bitumen. This pump is sucking the bitumen from a tanker through a
flexible hose. The problem is that after all bitumen is sucked from this
tanker, some of it remains unsuckable in the suction hose. This regularly
happens and it is a daily headache for the end user. My question is as follow:
why this pump can not make stripping after each delivery? Please reply by
return and advise how can we solve this problem.
THANKS & B.R
AHMED SOBHI
SALES MANAGER
ROTATING EQUIPMENT DIVISION
DAFF TRADING & OIL SERVICES EST.
POB 7399
Answer: Dear
Ahmed,
This is what I think is happening:
Pump people have an expression: “If you get the stuff to the pump,
- we can pump it!” – and there is some truth in it. Before a pump can move the
fluid out, the fluid must first get to it. The pump does not actually suck the
fluid, - what it does is moving the fluid out into discharge, creating space
for more fluid to get in. What makes the fluid to get in is suction pressure.
On the surface of your tank you probably have atmospheric pressure. The height
of the fluid adds to that pressure. As tank empties, the level drops, and thus
total suction pressure drops. There are friction losses in the suction pipe (or
hose) and suction pressure must overcome that. If suction pressure is not
enough to overcome theses losses – no pumping. The more viscous the fluid, the
more this would be an issue. Bitumen is typically very viscous.
Also,
as the level drops and gets to the point that the inlet port is exposed to the
air, it may be getting into the pipe, creating additional problem.
Couple
of things you may consider:
a)
Make sure the pump is as close to the tank as possible, i.e. make
the connecting line as short as you practically can
b)
Make sure the hose is not blocked, kinked or cut, allowing air to
get in somewhere along its length
c)
See if the pipe connected to the tank is at as low level as
possible – maybe take it from the very bottom?
d)
Perhaps use a larger hose, to reduce friction?
e)
Would it be possible to preheat the bitumen to make it less
viscous to flow better? (at least at the end)
I am also forwarding your
question to Mr. Tate Coghlan, who is a US Regional Manager for Monoflo Pumps, a manufacturer of Progressing Cavity pumps. They
may have additional suggestion, and could also help you locally via their
international office, if the problem still exists.
Let us
know if these suggestions helped.
Best
regards,
Dr. Lev
Nelik, P.E., Apics
Pump
Magazine & Pumping Machinery
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Question 32: Dear Dr. Pump, Happy New
Year!
Can you tell me more about the effect of temperature on NPSHR?
I know that result of test of NPSHR with warm water is less than with cold
water, but I would like to know, why? What are the technical reasons for this?
Best regards,
M. Moloudy
Answer: The cold water
is a much “tougher” with regard to cavitation as compared to warm water. The
explanation goes back to the fundamentals of thermodynamics of cavitation. The
vaporization (boiling) of liquid in the process of cavitation is a thermal
process, and depends on fluid properties, such as pressure, temperature, latent
heat of vaporization and specific heat. To make vapor form, the latent heat of
vaporization must be derived from the liquid flowing through the pump. This
flow of heat can only be possible when the liquid temperature is above the
saturation temperature at the main pressure in the low-pressure zone where
cavitation is about to begin. In other words, the pressure in the cavitation
region must fall below the saturation pressure corresponding to the liquid
temperature.
As we know, pump head begins to drop when cavitation begins – as
bubbles block the passages more and more. Keep in mind that the term “pump
head” ultimately means “energy per unit of mass flowing through the pump”. This
energy (i.e. enthalpy) is related to specific heat as:
Dhf = CL x DT
This heat transfers transforms some liquid into vapor. The ratio
of the resulting vapor volume to the remaining liquid volume would determine
the extent of blockage of the passage by vapor. “B” is a thermal
criterion, defined as:
B = (Vvapor
/ Vliquid) x (Dhf
/ L)=(Vvapor / Vliquid) x CL x DT
As you
can see, the “blockage of the impeller passage” is directly proportional to the
latent heat, i.e. more pronounced for cold water then warm water.
This affects
not only loss in performance, but also the damage to the pump. Vaporization
causes performance drop, but their eventual collapse (implosions), as they move
on to a higher-pressure zone, is likewise more violent for cold water, as
compared to warm – for the same reason, back to enthalpy and specific and
latent heat.
In
fact, the analogy can also be extended to hydrocarbons. As you know, API allows
NPSH corrections for hydrocarbons, versus tests on cold water. The reason –
hydrocarbons are less damaging from the cavitation standpoint, as compared to
water. In practice, however, this rule is not usually enforced, as most people
prefer to rather have some safe margin of NPSH, instead of cutting “too close
to the wire”.
I hope
this helps with your question. There is more to it, but is too technical to
cover here, although we would be glad to provide you with a more detailed
explanation if you should need. Please note that many of the NPSH-related
aspects are covered during the Pump School classes, and you are welcome to
check the schedule and the Agenda at the PUMP SCHOOL (click) section of the Pump
Magazine.
Regards,
Dr. Lev
Nelik, P.E., Apics
Pump
Magazine
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Question 33: Hi,
I would like to learn how the NPSH test of a pump is
performed. Would you please explain the method by an example?
Thanks,
A. Kiziltan
Answer: Dear Mr. Kiziltan,
Pump
manufacturers conduct NPSHR tests in one of two ways: either by throttling the
suction valve, or by vacuum tank.
a) Valve
throttling at several flows:
At each flow, as valve is throttled, the suction pressure at the
pump suction flange drops, i.e. available NPSH (NPSHA) gets lower and lower. At
some point, cavitation begins, and with it, pump Head begins to drop. When it
drops by 3%, the value of NPSHA there is what is called NPSH-required, i.e.
NPSHR.
At example above, values for NPSHR are obtained at (4) flows.
Then, these (4) values of NPSHR (the 3% note NPSHR3% is usually
omitted) are cross-plotted versus flow, which is what you see pump
manufacturers publish in their catalogs, called a pump NPSH-curve:
NPSHR
b) Another method is reducing suction pressure by
providing vacuum (called Vacuum Suppression Test) on the surface of the closed suction
tank. The rest of the procedure is the same as in (a).
Each methods has its benefits and drawbacks. Vacuum testing (b) is
more “pure”, since there is no disturbance in the inlet pipe by the throttled
valve, so for a more “scientific” experiment this may be preferred. The valve
method is simpler, and cruder so to speak. Depending on the type of a valve,
its position, etc., the results could be a little different. However, in
practice, pumps always have some sort of disturbance-causing equipment in line,
and thus represents a more real (“field”) situation. It will likely produce a
more conservative results, which is better from the standpoint of some extra
margin of NPSH.
For more information – please apply a SEARCH function on the
Home Page of Pump Magazine. There are several articles and questions posted,
which you may find helpful, regarding the NPSH topic.
Let us know if this helps,
Regards,
Dr. Lev Nelik, P.E.
Editor, Pump Magazine
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Question 34: Dear Dr. Pump
One of the consultants put a condition for sewage pump
suction, which is the velocity of the liquid at the eye of the impeller should
not exceed 3.7 m/sec, and that is for the submersible sewage pumps. Please advise
us the technical reasons for such condition?
In addition to that, we want to know how can this
parameter (eye velocity) be reduced, and what other factors could effect that?
Thanks and best regards,
AHMED SOBHI
SALES MANAGER
ROTATING EQUIPMENT DIVISION
DAFF TRADING&OIL SERVICES EST.
Answer: Your consultant
is right being concerned with velocity at the suction pipe. The higher the
velocity – the more are the hydraulic losses and less is static pressure that
remains at the pump inlet, which reduces NPSHAvailable. An approximate
rule-of-thumb number for suction velocity is 5-10 ft/sec (1.5 – 3 m/sec), which
is used to size the suction pipe.
Among other important parameters is the distance from the pump
inlet to the bottom of the sump, as well as distances from the pump and side
walls of the sump. Pump submergence is important also, to make sure there is no
entrapped air or vertexes that can get “sucked-in” and cause loss of
performance, vibration, and damage. To recommend exact numbers, we would need
to know more details, such as pump size, performance, geometry, NPSHA, sump
diagram, etc.
Another important factor for water, sewage and, especially,
marine, applications, is a choice of materials. These applications often have
problems by the combined attack of corrosive environment (for example salt
water or brine), as well as abrasion (entrapped sand, run-off debris, etc.),
plus cavitation. Bronze impellers are known to have problems, but even
stainless steel could be prone to short life. You may consider engineered
composites materials, such as Simsite, which is graphite fiber based composite
(it is not plastic), with strength similar to metal, and 3-5 longer life as
compared to bronze. Prime examples are marine applications, Navy, ship pumps,
and sewage treatment usage.
You can use SEARCH function on the front page of Pump Magazine to
read more about NPSH, cavitation, Simsite, and related topics.
I hope this helps,
Best regards,
Pump Magazine
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Question 35: Good afternoon Dr Pump,
I have read many of your papers and found them all very
informative and most useful. I wonder if you can help me out with some
ammunition. I have a chilled water circulation pump that destroys itself within
weeks of start-up. I have told the OEM that I want a pump with a more robust
construction and have found the L3/D4 number is 85 (US units) (in metric units
equivalent to 5). I have read somewhere that the industry standard is 60
(1/inches, English) but I can’t find any standard to back up my argument.
Can you help?
Regards,
Keith Townson
Senior Rotating Equipment Engineer
Qatar Petroleum
Answer:
Dear Keith,
Take a look at Article #4 of Pump Magazine, where there is more
discussion on L3/D4 criteria. The so-called
“L-cubed-over-D-to-the-fourth” is a criterion of a rotor stiffness, i.e. its
ability to resist deflection by the load. You are correct, single-stage
overhung impeller type pumps have this ratio vary between 20-120 – obviously,
the lower the better.
There is no standard actually, where the number is specified, but
the users, engineering firms, or contractors, sometimes specify the number,
when they want to ensure equipment is robust and reliability is particularly
important.
Do not overlook another very important factor. Shaft deflection
(y) is directly proportional to the load, overhung length
(“cubed”), and inversely to the shaft moment of inertia (where D
is in 4th power). That is how the “L3/D4” factor comes in. When we
compare the L3/D4 numbers, it is usually assumed that we are talking about the
same load. The weight (W) of the impeller adds to the load. If this weight is
reduced, the load is less, deflections are less, and life is longer. For many
applications (and, by the way, a previous reader’s question relates to this
too), - if the metallic impeller is replaced by a non-metallic impeller, the
effect on reliability improvement can be dramatic.
y = k x (W) x (L3/D4) –
i.e. “W” has as much effect as L3/D4 !
The problem is that plastic pumps impellers, while very good from
the corrosion standpoint, are not strong enough, and have limited temperature
capability. Structural composites, however, do not have these limitations. For
example, Simsite structural composites, manufactured by Sims Pump company, have
strength equal to metal, excellent resistance to corrosion, and superior
cavitation characteristics. It outlasts bronze by a factor of five,
which is why it is an excellent material for marine and navy applications, such
as seawater pumps, desalinization stations, water and waste treatment, where
their abrasive characteristic against sand and particulates makes them a choice
material. Simsite weighs only 20% of metal, which means dramatic
reduction in deflections and very significant equipment life improvement.
When a particular application has a problem, it is not easy to
change the L3D4 shaft ratio, but replacing the impeller is simple, and has an
immediate effect of 3-5 times life improvement. Of course, for new
installations, it makes sense to even get a complete pump made from the Simsite
material, especially if corrosion and abrasion are considerations.
I imagine in the Middle East, where your company is based, this
material could provide an excellent solution to numerous reliability problems
in the ocean water applications, salt water pumps, marine and navy pumps, as
well as water purification and desalinization projects. Utility plants, using
salt water for cooling pumps, as well as brine applications, are another
example.
You can do a SEARCH function on Sims material at
the front page of the Pump Magazine section.
Let us know if this information was helpful,
Pump Magazine
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Question 36: Dear Dr. –
I'm a new reader and I appreciate your magazine very much.
I've just read the Article #10 about Pump-Out Vanes and I'm one of that guys
that would like to spend nights without sleep reading something about new
things in pumps. As you wrote in the article I'd like to have more material
about that, so if you can get to me the material I'll appreciate it very much.
I have already a book by Stepanoff which is, I think, is
one of the most important ones, and Lobanoff too, but I didn't know about
Zanker and others. I appreciated really much your way of explaining complex
things in a simple language, to answer and clarify questions, without
sacrificing the technical part.
So please let me know.
Best regards
M. Meana
Answer: Thank you for
your kind compliments. Stepanoff’s book is truly an excellent source of information
for the pump designers. It covers a variety of pump topics, including axial
thrust. However, it does not cover fully how thrust varies with the variation
in the gap between the pump-out vanes and the stationary wall of the casing.
Nor does it cover the effect of the pump-out vane height. These could be very
significant. A plus (+)1000 pounds of axial thrust toward the
impeller eye could become a +1400 pounds, or even minus (-)600
pounds thrust in the opposite direction if the gap changes, or the vanes get
machined off, or worn out. In case of open impellers with adjustable front
clearance (impeller to casing), the reason and justification for the
adjustability feature is the ability to restore the gap, thus reducing the
front leakage and restore the efficiency back to the original “non-worn” value.
But, - the movement of the impeller forward also opens up the gap in the back!
While the back side of the impeller has little effect on efficiency, it does so
on axial thrust! New problems with bearings overload, seal leakage, etc. may,
or sometime do, pop up – all of a sudden, as far as a pump user sees it.
For general information on pumps types, equipment reliability
questions – from the pump users point of view, we recommend a book (click on) “Centrifugal
and Rotary Pumps: Fundamentals with Applications”, by CRC Press,
1999. Zanker’s paper, to which you are also referring, is essentially a
detailed research publication (about 15-20 pages), and deals with the pump-out
vanes and their effects in much greater detail. At this time, PUMP MAGAZINE is
working on putting together a CD with a complete set of Pump Course Notes (200
pages), which will be available for sale for $200 (shipping is free within the US;
add $10 for international). We could send you an advanced copy of the CD and
include Zanker’s paper on it if you would like.
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Question 37: Dear Dr. Pump,
I want to know the difference between pumps with API-610
and ISO standards.
Would you please explain the differences between pumps
manufactured based on API-610 and pumps per ISO standard per following points:
1. LIFE TIME OF
THE PUMPS
2. LIFE TIME OF
BEARINGS
3. VIBRATION
LIMITS
4. ALLOWABLE
NOZZLE LOADS
5. BALANCING
6. BEST
EFFICIENCY POINT LIMITATIONS
7. ALLOWABLE
DESIGN PRESSURE
8. MIN. MATERIAL
ALLOWANCE
9. SEALING
STANDARDS
10. BEARING OIL
LUBRICATION
11. MIN.
CORRESION ALLOWANCE
Best Regards,
H.R. Pourahmadian
Machinery Dept.
Answer:
Dear Mr. Pourahmadian,
There are several main standards covering pumps. They came about
at different times and by and/or for different markets and industries. Among
the most known are Hydraulic Institute, ANSI, ISO, API and PIP.
Hydraulic Institute. This is
mostly generic, and not as detailed. It deals with definitions, pumps
classification, types and nomenclature. Back in 1950s, when the US pump world
basically consisted of 5-6 major pump conglomerates (Ingersoll-Rand,
Worthington, Allis-Chalmers, Byron-Jackson, United, and a few others), the
Hydraulic Institute basically represented a “club” where top executives would
periodically meet, talk about business, relax and see what is going on. At that
time, it was (and still is) a very prestigious organization, with membership
fees to $15,000 or even more. Naturally, smaller companies did not belong.
As time went on, a pressure to produce more “tangibles”, and less
good-time-chat, started to shift the nature of the HI into having more
technical committees, with executives making room for engineers, who would
produce recommendations on things like sump design dimensions, pump
efficiencies, vibration levels, etc. The fees dropped, so that smaller
companies began to be able to rub shoulders with the “big guys”, at a “meager”
$3000 or so.
As a result, the HI is undergoing a transformation towards a more
practical, results-oriented organization. It will take some time, and the
present HI publications are still mainly useful for larger pump manufacturers
who deal with sophisticated pump users, and large engineering houses. HI is
putting a lot of effort into meetings and conferences to discuss and address
pumps reliability, equipment life extension, etc., which could become a
promising beginning to revitalization of the practical aspects and concerns of
the pump users community.
ANSI. This is an abbreviation for the American
National Standards Institute, published by the ASME (American Society of
Mechanical Engineers). It has several sections: for horizontal end suction
centrifugal pumps; another one for vertical in-line centrifugal pumps;
seal-less mag-drives for chemical processes. These have more specific data,
regarding standard pump sizes, to make pumps made by different manufacturers
interchangeable for a given installation. They do not impose any specific
dimensional standards on the pump internals (except for the seal chamber
dimensions). For example, impeller width, or casing-to-stuffing-box fit is
entirely up to individual manufacturer.
API. Issued by American Petroleum
Institute, API-610 is for centrifugal pumps. While ANSI mainly addresses pumps
interchangeability, the API addresses the “robustness” of the design, to ensure
trouble free operation for the demanding and tough environment of the oil
refineries. It specifies whether an impeller should or should not have wear
rings, determines allowable nozzle loads, imposes rotating parts balance
methods and criteria, specifies piping plans for specific service, addresses
lubrication, bearings and baseplates, as well as materials of construction.
These standards are available for purchase from the API in Washington, DC and
the number is 202-682-8000.
As environmental pressures increase, the attention to seals
increases also. There is now an API-682 standard, and excellent publication,
dealing with seal dimensions, allowable run-outs, plans, barrier fluid, as well
as has sample calculations for heat generated by the seals.
API-676 is similar to API-610, - but for Positive Displacement
Rotary pumps. Ironically, there is not a single rotary pump to this day (to our
knowledge) that complies fully with the API-676. Some come closer
then others, and many claim conformity to the “intent” of the API-676, but none
claim full compliance, “to the letter”, contrary to what API-610 pump
manufacturers do.
API-685 is for the seal-less centrifugal pumps, for petroleum,
heavy-duty chemical, and gas industry services. It is still a relatively
unknown publication, with few seal-less pumps within the refinery applications,
due to high temperature service, not easily handled by the magnets of the
seal-less pumps.
PIP. About 10 years ago or so, several major
oil companies, such as Shell, Amoco, Aramco, Phillips, as well as some chemical
companies, such as DuPont, Eastman, Celanese, etc. felt that the API-610 was
not sufficiently conservative, and too influenced by the interests of the pump
manufacturers. The issue of pump baseplates, for example, was a sore point of
contention between the pump users, who felt the baseplates, overall, are “too
flimsy”, and the pump manufacturers, who felt that the added cost to make the
“more robust” baseplates was not justifiable, and an unnecessarily requirement
by the users. In response to that, the users formed their own specification
(Process Industry Practice), as recommended Practices for Machinery
Installation and Installation Design. This standard reflects a lot of details
of piping, baseplates, grouting, etc. – clearly a users view on reliability requirement.
PIP, however, did not find much actual implementation, and sort of
stayed dormant as a silent reminder to the manufacturers “not to question” the
users reasons for more robust equipment, and just comply. Commercial and competitive
market forces took care of this issue, and the need for the PIP became less
critical. Today, API-610 is a most typically used for the refinery
applications.
ISO. There are several ISO specs, developed
in Europe, and the main one is the ANSI-counterpart, specifying outside
dimensional envelope of the “metric” pumps for chemical services. For
refineries, however, API-610 has been and continues to be an internationally
accepted standard. A typical API-610 centrifugal pump is a much more robust and
engineered design as compared to an ANSI, and probably 50-100% more expensive
as well.
The ISO 13709 intends to formally accept and adapt the API-610 as
is, thus making it an ISO Standard for the refinery pumps, just as API-610.
Therefore, all of the (11) points that you listed are addressed
within the API-610, which now (as of July 1, 2003) corresponds to the ISO
13709. The joint working group (JWG)
will be meeting In October 2003 in Houston, to work on improvements to ISO
13709 with the goal of republishing it within a couple of years to replace the
new API 610 9th Edition, and, at that time, have API adopt back the
next ISO 13709 publication as API 610, 10th Edition, probably in the
2005-2006 time frame.
We are not elaborating on each specific point that you listed, as
that would require an extensive and lengthy tractate on a Standard itself,
which is accomplished best by actually reading the API-610 Standard itself. We
do that during our PUMP SCHOOL Training Sessions at PUMP SCHOOL.
Thank you for your question. We welcome our readers to comment,
and will be periodically updating this interesting and important section on
API/ISO/ANSI/PIP/HI Standards, as we get more feedback from our readership –
whose comments are welcome.
Editors,
Pump Magazine
Dear Dr. Pump,
Many thanks for your useful reply. I have also another
question. Could you please explain what happens if a pump starts with fully
open valve (end of the curve)? Also, to avoid the disadvantages of starting
pumps at above-mentioned condition, what can be done for starting the stand-by
pumps when they are in auto-start mode?
Best regards,
H.R. Pourahmadian
Dear Mr.
Pourahmadian,
For pumps with low to medium specific speed (NS < 3000), which
is the majority of pumps at chemicals plants and refineries, etc., the pump
horsepower rises with flow. The lowest horsepower is near the shut valve
condition. This is why it is best to start pumps with discharge valve slightly
cranked open, and open it up more, to the desired flow, after a pump has been
started. This puts less stress and in-rush current on the motor, and it will
last longer.
For high specific speed pumps, the shape of the power curve is
different, and the lowest power may not be at the shut valve, i.e. starting of
those is not as simple.
Regarding the auto-start mode, - we would like our readers, among
the pump users, to comment on their practices, and will publish their comments,
so that other users may exchange their views and experience with this.
We hope this helps,
Pump Magazine
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Question 38: Dear Dr. Pump,
I am in the midst of an engine
ancillary systems design, part of which is a water pump. The engine is for
racing, so high efficiency is a primary requirement. It has a flow of 320
litres/minute at 48m head and pump rpm of 5843. This is a point on the system
curve (flow through an orifice), which I have taken as the BEP as the pump will
run primarily at this
speed. This is my first centrifugal pump design and I am
calculating the various physical sizes of the pump from a collection of texts,
but there are still areas for which I have little detail or explanation.
Reading the Q&A Section of Pump Magazine has been very helpful. I was using
Stepanoff method, but not from his book, but a "paraphrased" by
another author. Reading the ratios for dimensions off a chart makes me a little
nervous. I having doubts as to whether this is the best route for a pump of
this type, which has so much better manufacturing methods, smaller clearances,
fewer operating hours, and better operating environment a compared with
industrial pumps. Basically, faced with
a clean sheet and a tight deadline, what would be the most practical reference
book to get a good design done? (Preferably with metric units).
Thanks in advance,
Philip le Roux
Answer:
Dear Philip,
You are tapping on a “heart of a pump design” – pump hydraulics.
It is rather involved subject, much of which was developed over many years of
trial-and-error by the pump hydraulics engineers, books, tests, field
experience, etc. Stepanoff book is a good one to start, but intended mainly for
a “specialized audience” of pump designers. There are many other books on
pumps, as well as technical papers, conference proceedings, including, for
example a book recommended by Pump Magazine, “Centrifugal and Rotary Pumps”, as
shown in PUMP SCHOOL Section, which contains hydraulics theory, with examples
and applications. Unfortunately, pump hydraulics is not an area that can be
learned quickly, even after reading a book, as much of it also relates to a
“non-hydraulic” side of a pump – its dimensional constraints, weight and space
limitations, the shape of the performance curve (e.g. it is possible to “force”
more head at the BEP, but end up with a “drooping curve”, - a known problem
when pumps operate in parallel).
Using ratios actually is not a bad way to begin the design. Both
Stepanoff and Andersen – the two “Classics” of the pump world, relied on these
ratios heavily, although from the entirely different perspectives, which
resulted in two schools of thought in hydraulics: Stepanoff’s is of blade
angles, and Andersen using area ratios method.
In your example, the first thing either one of them would do is
determine pump Specific Speed:
NS =
5843 x 851/2 / 1573/4 (I converted flow to GPM units, and
head to feet) = 1214
(Read more on Pump Specific Speed in other sections of
PUMP MAGAZINE, using SEARCH function on front page).
This leads to a selection of a so-called head coefficient,
and the impeller diameter (OD). In your case, the impeller diameter would be
approximately 3.9 inches (almost 100 mm). The exit width would be approximately
0.36” (9 mm).
Then, the impeller eye size would be determined, the number of
blades, their angles, etc.
Depending on the rest of the design details, an open or closed
impeller type would need to be decided on. And, the choice of the material is
critical. In your case – pumping water is not as bad as, say, sulfuric acid –
from the corrosion standpoint, - but, the speed (RPM=5843) is rather high. The
higher the speed – the shorter the pump life: wear is normally estimated to be
a function of RPM - cubed! (RPM3).
Certain materials resist wear, abrasion, as well as cavitation damage better
then others, and this is very important consideration.
A final impeller layout would need to be accompanied by the
cross-sectional layout of the pump, to show how all components fit together –
impeller, volute (or diffusor), casing, seal, etc.
PUMPING MACHINERY offers such Consulting and Specialty
Pump Design services, and would be glad to assist. If you are interested, email
or send us pertinent information, and we can produce a hydraulic design of your
centrifugal pump, as well as recommend manufacturers who could produce a
complete impeller, made from proper materials, as well as related parts (casing,
bushings, rings, etc). To start, you may want to fill out the Form in Section CONSULTING: HYDRAULICS, DESIGN AND APPLICATIONS.
Sincerely,
Dr. L. Nelik, P.E., Apics
Pump Magazine & Pumping Machinery
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Question 39: Dear Dr. Pump,
Can you say something about the minimum continuous flow
that must be guarantee to a centrifugal pump? And how it's related with the
viscosity?
Best Regards
M. Meana
Answer:
Dear Mr. Meana,
Hydraulic Institute addresses the general concept regarding the
“..minimum and maximum flow, at which they should be operated continuously of
for an extended period of time..”. The Standard continues to say that
“..Operation of pumps at reduced capacities may lead to the following problems:
temperature buildup, excessive radial thrust, suction recirculation, discharge
recirculation, insufficient NPSH..”
Regarding viscosity, the usual pumping limitations of centrifugal
pumps apply (see more discussions on viscosity via SEARCH function in several
sections of PUMP MAGAZINE). Since pump energy level is an important
consideration for the minimum flow, as the HI states, the corrections for flow,
head, efficiency and power thus affect the MCSF.
PUMPING MACHINERY can perform specific calculations for the
temperature rise, to determine the thermal limitations on the pump flow, as
well as calculate the radial thrust as various capacities, and produce the
report with recommendations. If you would like us to perform such work, we
would be glad to assist with consulting for your specific application. We would
need to know the pump type, size, and other parameters, as stated in Section CONSULTING: HYDRAULICS, DESIGN AND APPLICATIONS.
Sincerely,
Dr. L. Nelik, P.E., Apics
Pumping Machinery
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Question 40: Dear Sir,
We are thinking of using an inducer 4" OD. Are
there any general tech papers out there?
How are they calculated for flow?
Duane Leonhardt
Answer:
Dear Duane,
Inducers are certainly a “pump art of its own”! Hydraulically,
they are essentially very high Specific Speed axial flow impellers, with Ns =
20,000 – 25,000. They generate very little differential pressure (which is why
Ns is so high), just enough to boost up the inlet pressure to the inlet of a
“normal” impeller that follows the inducer. Hydraulic blade loading is so
small, that they operate with a cloud of vapor trailing along the suction side.
There are papers written on this subject, but they are really not general, but
highly technical, and hydraulically involved. ASME Transactions could be a good
source for information if you are interested.
I am not sure about your question on calculations regarding flow –
if you mean how they are designed, then the procedure starts off very similar
to any other high-specific-speed designs, with a number of blades usually 2-3,
and blade angles set to produce very little head. However, the shape of the
blades, and the way they unwrap along the passage, is more critical then in
case of lower Specific-Speed machines: the loading must be more gradual, to
keep uniform transformation of energy. They also can be more susceptible to
instabilities. For example, low flow instabilities, and the radial thrust that
results, can “drive the inducer” into a
shroud wall of a casing, taking out the clearance, and causing a mechanical
contacts and failure.
We would be glad to perform a hydraulic design for you, or
evaluate the existing design, with recommendations.
Sincerely,
Dr. Lev Nelik. P.E.
Pump Magazine
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Question 41: Dear Sir,
Thank you for this useful site. Would you please explain
the difference between using single speed and variable speed pumps in terms of
performance, NPSH curves and what precautions should be taken when using
multi-speed pumps?
Thank you,
Marwanco Company, UAE
Answer:
The main advantage of the variable speed drive (AC or DC controllers) is that
when a pump flow is changed by the speed of the motor, the relative pump flow
(in proportion to the BEP) does not change:
