PUMP MAGAZINE: Questions and Answers
(81-90)
Question #81 Dear Doctor Pump,
I found your website during
search on the Internet and I am very thankful to you for the assistance you are
providing regarding the pump problems. I have two questions regarding the pump
selection and design parameters:
1. How does vapor pressure of
liquid affect the selection of a pump?
2. Our company is in the
design stage of the pumps, and I want to know how to calculate mass moment of
inertia of impeller. Normally, in
Thank you,
Pankaj Lawate
Answer:
Dear Pankaj:
Regarding your first question: vapor pressure is one
of the subtractive terms in calculation of NPSHA (net positive suction head),
and is of great importance. You may read more on this at our Pump Magazine via
Search function (type NPSH, and see many subjects pop up).
Regarding the moment of inertia (in the US, it is
typically expressed as WR2, where W is weight in pounds, and R is
radius in feet). This parameter is used in water hammer analysis, pump runaway
speed (reverse rotation during check valve failure), and similar transient
(non-steady-state) hydraulic problems.
Normally, a complete pump rotor is analyzed, and often
even a complete train, consisting of all rotors coupled together: pump,
coupling, motor, etc. Impeller is one of the main components with rather
significant value of WR2, but other components need to be considered
also (coupling, for example, may have a very significant effect on
rotordynamics).
For the impeller, a pump manufacturer can usually
supply the value of WR2, but, for a quick approximation, you can
assume the impeller as a disk with same weight, and same outside diameter, but
the width adjusted to correlate with the known impeller weight. The radius is
then the same as for a disk, and you can get it from any text book on dynamics.
Best regards,
Dr. Lev Nelik, P.E.
Pumping Machinery
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Question #82 Dear Dr. Pump,
We are Electrical Engineers
with pumping applications in Oil & Gas production, and we receive your good
and valuable newsletters.
I have a question
as to what type of pumps are referred to as RIVER PUMPS in Oilfield
applications?
Your kind guidance
and advise is highly appreciated.
With many thanks,
Tom Moody
THE CONCORD GROUP
Sherman Oaks, CA 91403
Answer:
Dear Tom,
I must admit – I have not heard this term, and am
hesitant to speculate as to what I think these are. However, we will post your
question on our Q&A Board, and if we receive any feedback from our readers
– we will let your know.
Dr. Lev Nelik, P.E.
Pumping Machinery
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Question #83 Dear Dr. Pump,
I have a question on pumps hydraulics. I
have the performance curves of one of our pumps running at 2300 rpm and I want
to generate the curves at other speeds. Am I right if I use the affinity laws
to get Q, ∆H
& BHP and then calculate their respective Efficiency by using the following
formula:
Eff = [∆H
(ft)* Q (gpm)* S.G.] / [3960 * BHP (hp)]
Please help me. I tried to calculate the Eff
using the above formula for 2300 rpm and then compare it to the curve Eff but
they were different. What did I do wrong?
Your help will be highly appreciated, best
regards.
ABDALLAH AL-GHAMDI.
Aramco Company,
Pipelines Dept./PTSD/PSU
Rotating Equipment
Answer:
Dear Abdallah,
Your calculation of efficiency is correct.
Just make sure you calculated the performance curve at various speeds
correctly. Flow varies directly with RPM; head as a square function; and power
as cube. Pick a point on the original H-Q curve at 2300 rpm and apply affinity
laws to them. Then pick another point and do the same. Do that for 4 or 5
points, and then plot them. You will get a new H-Q curve at new speed. Do same
for power. Then calculate efficiency at several points, and plot the efficiency
curve.
The resultant curve will have a BEP
(best efficiency point) shifted to the left (if new speed is lower), or to the
right if it is greater. As you found out, efficiencies are different, of
course, at same flows, but – at the BEPs – it is the same.
As a minor note, technically
speaking, the BEP efficiency is not exactly the same, though, - although not
because of the affinity laws, but for a different reason. At slower speed, the
proportion of the hydraulic losses is a little greater in relation to the
overall power, as compared to what happens at higher speeds. Hydraulically
speaking, it has to do with a so-called Reynolds number, but the overall effect
is relatively small for small changes in speed. For example, a BEP efficiency
of 80% at 3000 rpm may become a 78% or so, at 1500 rpm.
I hope it helps!
Dr. L. Nelik, P.E.
Pumping Machinery
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Question #84 Dear Dr. Pump,
I seek your expertise
advice on selection and application suitability of oil mist lubrication
system for centrifugal pumps operating in elevated temperature service 200 deg.
C and above:
a) Is oil mist lubrication successful for above application and is there a
proven record over conventional oil lubrication?
b) Any information on the organization where
this system operates successfully for such high temperature service for
rotating equipments
c) Which type of oil mist lubrication is
most successful: pure mist or purge mist?
d) Does this system really eliminate (or
reduces significantly) bearing failures, as claimed by the supplier? Is there
any statistical data available to examine this?
Kindly guide us to get information on the
subject oil mist lubrication system.
Regards
Sourav Kumar Chatterjee
Manager Rotang Equipment
HPCL Mumbai Refinery
We have asked Heinz Block, who is a
renowned authority on the subject of lubrication, to comment.
Since there are various ways and reasons to apply
lubrication systems, the gist of Heinz’ recommendation for the reader was to first
obtain basic knowledge of the principles of bearing lubrication. It is
important to first understand the basics of bearing operation, and fundamentals
of various oil lubrication systems, as there are many variations of these.
Heinz expressed some doubt if the reader’s bearings indeed operate at a
temperature of almost 400 degrees F, and if so, it would be of interest to know
why, and what is the application specifics. It is premature to discuss oil mist
lubrication for utterly hypothetical applications. What is a temperature of the
pumpage itself? If it is hot, would it be better to have a bearing housing
design that isolates it (thermally) from the liquid (hot) end, or using cooling
coils, or similar methods of keeping the whole bearing housing (and thus the
bearings) cool? If, on the other hand, the pumped liquid is not hot, then the
reason for hot bearings could be pipe strain, misalignment, or numerous other
problems. If so, these problems would have to be addressed and resolved first.
It might be possible that, by taking care of the problem properly, the whole
issue of hot bearings might not be an issue after all. Only then, when all
factors that may be affecting bearings operation are understood and are taken
care of, the lubrication method itself should be addressed.
Learning the fundamentals, to compare pros and cons of
various systems should be the first step. As a suggestion for a source of such
information, please refer to Lube Systems Company, or to the Bloch/Shamim text
"Oil Mist Lubrication: Practical Applications", which can be
purchased from Amazon.com.
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Question #85 Dear Dr. Pump,
Is boiler feed pump recirculation valve
leak-through - an industry wide problem?
I'm an Operations shift supervisor at
Hawaiian Electric Co's Kahe Power Station. Kahe consists of six conventional
boiler-steam turbine-generator units ranging from 90 to 142 MW. We maintain our
boiler drum pressure between 1800 and 1890 PSI. Each unit is equipped with two
boiler feed pumps. Our units are not equipped with Deareators, therefore our
boiler feed pump recirculation valves discharge into the main condenser. Our
units cycle to low loads at night so we usually take one boiler feed pump off
from each unit every night. Hence, the boiler feed pump recircs’ stroke open
when we take the pump off at night and stroke close when we put the pump back
on in the morning. Nearly all of our boiler feed pump recirc valves leak thru.
Our engineers have purchased a number of different designs with little success.
Within a year, they all seem to leak thru badly. Was wondering if this is an
industry problem or are we just using the wrong type of valve.
I have not had much involvement to date in
the company's effort to solve this problem. So I don't really know what sort of
valves we currently have in service. What I see on the web, though, suggests
that we ought to be using a stacked disk or drilled hole trim. Do you have a
suggestion?
Came across your column while surfin' the
web. Thanks for the help in advance!
Robert M. Clark, P.E.
Shift Supervisor
Kahe Power Plant
Hawaiian Electric Co.
Answer: We have asked our colleagues in the power generation to
comment. A general consensus was that the problem you are describing is rather
common. High differential pressure across the valve causes high velocities, and
thus erosion, flashing, etc. are the potential problems. This is what Maria
Branco of Keyspan Energy in Long Island, NY, and Jack Steenbeek of Consolidated
Edison, New York, commented:
Maria Branco:
Please tell your contact in Hawaii that I will be glad to come and help him -
if he pays for the trip of course! BFP recirc leak-through is extremely
common. The pressure drop across the valve trim is often greater than
2000 psi and the condenser vacuum draws the flow at high velocities. It's
called severe service. There are many valve designs that cut down the
pressure in stages to mitigate the destructive effect of the large pressure
drop. Examples are: CCI, Target Rock and Fisher have the labyrinth trim;
Yarway has multiple seats in series; Copes Vulcan has a hard seat and a soft
seat; etc.
All these valves will be leak-free for about 5 years: less if they are subject
to many openings and closings; more if they remain closed for long periods of
time. Their longevity can be improved by adding an orifice downstream of the
valve [closest to the condenser is best] to offer some resistance to
flow. However, the orifice will become the sacrificial component and will
have to be inspected with some regularity and changed out when it shows
sufficient erosion to render it useless.
At the various
KeySpan power stations the most common valves used in "on-and-off"
BFP recirculation are Fisher and Yarway.
Maria
Branco
KeySpan Gen Ops - Far Rockaway PS
Jack Steenbeek:
Our
system operates with BFP pressures of 2300 - 2600 psi and a back pressure of
80-100 psi. We use CCI stacked disc recirc valves with a class V shut-off in
our BFP recirc system and have had problems but they were erosion and
cavitation related.
Our
inlet piping was sized too small. We used to reduce inlet piping from 6"
to 2" which caused water velocities to increase by a factor of 9. This
created severe erosion in the carbon steel valve body, causing leak
thru of the valve and eventually leak thru of the outlet stop valve. We have
now switched to 6" inlet piping, WC9 valve bodies, class V shut-off,
a break-down orifice downstream of the recirc valve and at least 8 pipe
diameters of straight pipe downstream of the valve to allow for proper mixing
of the possible steam/water mixture.
In
the case of the problem described, I would say that they see a large
amount of flashing steam downstream of the valve. It is not clear however, if
the recirc valve is closed after the pumps are shut down. If one pump is
shut down at night I recommend that the valving to the condenser for this pump
is closed to obtain a good isolation. If the recirc valve does not close
completely you will see a continuous flashing of steam which will eventually
lead to leak thru conditions.
Jack
Steenbeek
Con
East
River Station
I would like
to thank Maria and Jack for their input, and we hope Robert and his colleagues
at Kahe station will find it helpful. Let us know.
P.S. By the way, Robert –
can Jack and I join Maria for a “pump troubleshooting visit” to Hawaii? Aloha!
Pumping Machinery
Pump Training, Consulting
and Troubleshooting
Question #86 Dear Dr. Pump,
I am finding it difficult to understand how
NPSHR as specified by pump manufacturers, which is tested on cold water, is to
be adjusted in terms of actual pumping liquid. As I understand it, NPSHR is the
minimum head of liquid column needed to suppress vaporization. Should NPSHR of
water and NPSHR for the actual liquid in operation be the same?
I have a problem. My pump NPSHA is 2.1m, and
my pumping liquid is Butane mixture at 68 Deg C and density of 488 kg/m3. The
NPSHR as stated by the pump vendor is 0.6m (based on water). Should I convert
NPSHR to my actual fluid first (BUTANE SG=0.45, vapor pressure = 13 BAR), to
get the adjusted NPSHR for Butane of 0.6/0.488=1.22m ?
So, there should not be any cavitation,
right? But actually the pump is cavitating! Why?
Regards
Manik
Answer: Dear
Manik:
It does not matter if the 2m level in the tank is
water, LPG, mercury, or any other fluid. True, 2m of water would have more
pressure at the bottom of the vessel then LPG, and 2m of mercury would have
much more pressure at the bottom of the vessel as compared to water. The
pressure would be proportional to the specific gravity. However, a definition
of “head” is “energy per unit of mass” – thus
mercury has more “energy” (more pressure) but also proportionally more mass.
The ratio of that energy to mass is the same.
As a reminder of basics, in US units:
Head
= Pressure x 2.31 / SG
For example, if you convert the 10 feet of butane
(SG=0.45) to pressure, you get:
P = 10 x
0.45 x 2.31 = 1.9 psi
Same 10 feet of mercury would have:
10 x 13.1 x
2.31 = 56.7 psi
So, either one, converted back to “Head” will end up
at:
1.9 x 2.31 /
0.45 = 56.7 x 2.31 / 13.1 = 10 feet
NPSHA, NPSHR, pump head, suction head, discharge head
– all are subject to the same relationship of energy-per-unit-of-mass, and thus
the NPSHmargin is likewise.
The reason for hydrocarbon correction factors has
nothing to do with this. They will cavitate by same laws as does water or any
other liquid. However, cavitating hydrocarbons simply do not do the same damage
to the pump as cold water does, and thus the correction is allowed to reflect
that.
Keep in mind a definition of NPSHR. That is such NPSHA
at which a pump has already lost 3% of it’s developed head – that is by
definition of the Hydraulic Institute. It starts to cavitate way before that:
first, the incipient cavitation, then more bubbles, and, eventually, 3% head
drop, - and then more. So, you are not correct in your understanding that the
printed values of NPSHR are no-cavitation values.
NPSHR is indeed tested on water, and is plotted as
such. However, some fluids tend to be less damaging on pumps as compared to others.
This is why, hydrocarbons, for example, have a so-called hydrocarbon correction
factors, which essentially allows less NPSHA for them as would be for water
(colds water is very damaging on pump inlets if in cavitation). Using the HC
correction chart, you locate the particular hydrocarbon on the chart at its
given value of temperature and vapor pressure, and read-off the correction
(substation) off the NPSHR. In no case, however, the note with a chart states,
this subtraction would be less then 50% of the original value.
The issue is not that a pump will not have
vaporization on hydrocarbon – but whether that vaporization would do
significant damage or not. Hydrocarbons do not have the same damaging effect,
at cavitation, as cold water does.
At any rate, I do not think it is a good idea to allow
the pumps into NPSH problem zone, even relying on corrections. Instead, improve
the situation by either providing more NPSHA, or select a pump with less NPSHR,
such as slower speed, better inlets design, etc. Even if cavitation damage is
less for hydrocarbons, pumps operating at cavitation, would have other
problems: vibration, pulsations, noise, thrusts, - all ultimately reducing pump
life. Many companies require at least 5 feet margin between NPSHA and NPSHR, to
make sure the pump stays away from the “nasty” zone, as much as possible.
In your case, NPSHA = 2m, and NPSHR = 0.6m for water,
- which would be even less if adjusted for hydrocarbon, - i.e. your pump should
not be cavitating. If you hear your pump cavitating and also experience
cavitation damage on impeller inlets and possibly on casing, may indicate the
losses are higher then you calculate, - perhaps there is an obstruction in the
inlet pipe, pipe reduction right before the pump inlet, etc. Otherwise, the
2-0.6 = 1.4m margin should be very sufficient to make sure there is cavitation
problems. So take another look at the pump inlet, and definitions.
You can get more info on NPSH on our web, by using
SEARCH function by keywords. I also recommend you consider our Pump School
training, which could be arranged for a group of your colleagues and you. Let
me know if this would be of interest.
Regards,
Lev Nelik, Ph.D., P.E.
Pumping Machinery
Question #87 Dear Sir,
Is API 610 Standard applicable for Cryogenic
Service (LNG or LPG etc)? If it is applicable, to what temperature? If is not
applicable, which standard to be followed?
Thank you and regards,
V. Srikanth
IDPE Limited
Answer:
API 610 does not specifically cover or not cover pumps used in cryogenic
services. Much of the information
regarding pumps presented in API 610 can be applied as well to pumps in
cryogenic services as to any other petroleum, petrochemical or gas industry
service; the primary difference being the application and service temperature
since cryogenic is generally taken to indicate temperatures colder than minus
73 C (minus 100 F). The critical areas
to be specially considered for low temperature services are materials of
construction, sealing and bearings / lubrication.
I am not familiar with an
international standard for pumping equipment specifically addressing cryogenic
services, although there are many references that can be used to assist in the
application and specification of pumps for low temperature services.
I hope it
helps,
Dr. Lev
Nelik, P.E.
Pump
Magazine
Question #88
During a pump performance test the overall
vibration of a centrifugal pump rotor met the vibration limits of
Table 2-5 of API 610 8th Edition. However, the discrete vibration component at
running speed is exceeded the limit of 75% of the overall limit as specified in
API. (Note: the rotor was balanced and witnessed prior to installation into the
pump and the out-of-balance was well within acceptable limits). The pump casing
vibration was also well within the specified limit. Can you advise what
potential impact, if any, the excessive running speed component of
vibration would have on the long term operation of the pump and is this a real
issue to be concerned about?
The pump Supplier is recommending we accept the
pump mechanical performance without further investigation. As I do not
understand the background to the above API requirement I am hesitant to accept
the performance of the pump. Note that the 9th Edition of API 610 has dropped
the requirement of the 75% limit at speed greater than or equal to running
speed. It now only specifies limits relating to sub-synchronous discrete
vibration, i.e. frequencies lower than running speed. Hence if the pump
had been purchased against the later Edition we would not be having a problem
to resolve.
For information the pump has a rated
capacity of 1177 m3/hr at a differential head of 1219 m. The
absorbed power is 3.85 mW.
I would appreciate any assistance or advice
to help resolve this issue.
Thanks
Neal
Answer: This
sounds like a test stand performance issue. If so, it could indeed have
been the result of coupling unbalance, misalignment, or test stand resonance.
If the only problem is the 1X component, and the overall is well within
specification, it should be satisfactory when installed in the field. For
assurance, the vibration should be monitored to be sure this is the case, and
the supplier should be held responsible for corrective action if it is not
satisfactory at final installation and operation.
We will forward to you any additional
comments from the readers for feedback.
Regards,
Dr. Lev Nelik, P.E.
Pumping Machinery
Question #89:Dear
Dr. Pump.
I have heard that API 610 Ninth Edition
"allows" user to employ non-API-style pumps even in hazardous and
flammable which can comply to ISO 5199 or ANSI/ASME B73.1. The problem is: I
could not find above statement on the standard. I am wondering perhaps the
"rumors" quoted from ISO 13709 which is said equal to API 610 9th
Edition?
The point is how equivalent ISO 13709 to API
610 9th Edition and how relax they are
to allow user to use ANSI/ASME B73.1 for flammable/hazardous services.
Regards,
Edi Hamdi
Mechanical
PT. Rekayasa Industries
Answer: API 610
never "allowed" users to employ "non-API style pumps for
flammable or hazardous services. In the
8th edition, 610 contained the paragraph (1.1.3) that advised "For
nonflammable, non hazardous services not exceeding certain ratings (which, not
by coincidence, lined up with the general ratings of ASME B73.1 pumps,
purchasers MIGHT wish to consider pumps that do not comply with API Standard
610".
In API 610, 9th edition, and in ISO 13709, which is
essentially API 610, 9th edition with minor ISO editing changes - no technical
changes, this clause has been removed.
It was felt that it was not the function of API to advise users when
alternative pump designs and standards MIGHT be acceptable; that is a user's or
purchaser's decision to make, based on many factors.
There are several references in our web site regarding
API spec, you can find them via typing a
"API" in a Search button
function on the entry page of technology section of our web.
I hope this helps.
Lev Nelik
Pump Magazine
Question #90:
I have an existing lift station on one of my
jobs with a 1/2 hp Vortex Ebara (Model # 50DWX) pump. My question is - how long of a distance can I
pump through a 2" PVC pipe to a gravity Sanitary Sewer line?
Thank you for your prompt attention.
Christopher C. Roberts, P.E.
Project Engineer
WilsonMiller, Inc.
Answer: We looked
up the pump performance curve from Ebara web site www.PumpsEbara.com, which shows these
pumps look like this:
Then, we also looked up the
performance curves below, to see what flow and head characteristics correspond
to these pumps. The vortex impeller construction, shown on the left, seems to correspond
to the curve 500WXU6.4 ˝ HP. At, say, 30 gpm, the head is roughly 20 feet, and
the motor speed is 3600 rpm.
The 20 feet does not mean the distance, nor elevation,
but a total head, which is, in general, a total of a static heat (lift), plus
friction losses in a pipe to a point of delivery. I assume your pump has to
lift the sewage some height first, and, since you did not indicate what it is,
let’s assume it is about 6 feet lift. Then, you have 20-6 = 14 feet “to spare”
on friction.
Assuming your pipe is 2” all the way, the velocity in
it, for 30 gpm, is 3 ft/sec (do the math to prove). Then, by solving the head
loss equation, or looking up friction tables for the 2” pipe, the length is
1057 feet, - assuming clean and smooth pipe. Conservatively, assuming there are
some turns, elbows, fitting, etc., subtract some of that, to be safe, to
perhaps around 800 feet or so.
Also, since you are using a “grinder” style impeller,
I assume your pumpage contains solids and other rough material. If so, the
friction losses would be probably greater, thus effectively reducing the
allowable length of pipe even more – and you would probably end up at something
closer to 500 feet or so.
I hope it helps. For more information, I recommend
your visit Ebara web site, - they have more information of similar nature which
you may find helpful.
Let us know how it worked out for you.
Regards,
Editor
Pump Magazine
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