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 India, pump manufacturers call it as a GD2 (G X D Square) parameter. Will you please tell me the formula for this or any suitable method to determine this GD2 value of a pump?


Thank you,


Pankaj Lawate

Engineer, India


      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




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
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




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.



Aramco Company, Saudi Arabia

Pipelines Dept./PTSD/PSU

Rotating Equipment Eng.


      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




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.
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.




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 Edison

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!


Lev Nelik

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?





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.




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.





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.



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.




Edi Hamdi

Mechanical Eng.

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.



Lev Nelik


Pump Magazine






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