Tech Quest

Field Problems
Design of Cooling water system

I am having a query about our system. Details are follows.

Suction Head of CW pumps 2.980 mwc

No. of CW pumps 2 (in Parallel Operation)

Discharge Pressure 1.7 Kg/cm2
Required CW flow to Condenser 5810 m3/hr

ACW Suction Pressure 1.6 Kg/cm2

(Tapping from CW header)

ACW discharge Pressure 4.4 Kg/cm2

CW Pumps' flow is calculated as follows.

Net Head developed (1.7 - 0.298) = 1.402 Kg/cm2

According to the Pump Curve referred at this pressure and at a Frequency of 50 Hz, flow is found to be 4880 m3/hr for each pump.

ACW flow is calculated in the same method.

Net Head developed (4.4-1.6) = 2.8Kg/cm2

According to ACW pump curve, flow found is 675 m3/hr

So, Total CW flow required = (CW for Condenser
+ CW for Aux. System)

= 5810 + 675 = 6485 m3/hr.

Flow that can be developed by both the CW Pumps, according to the curve, proportional to the Net Head developed is (4880 X 2 =9760 m3/hr)

At present, we are running the CW Pumps by throttling their discharge valves to 50 % and all the three Riser Valves opened by 35% each, to control the flow.

CW Pump Design Parameters

Discharge 3450 m3/hr, Head 20 M Speed 985 rpm NPSH 6.07 M Motor Capacity 250 kW

ACW Pump Design Parameters

Discharge 600 m3/hr, Head 30 M, NPSH 5.23 M Motor Capacity 75 kW Efficiency 81 %

In this condition how can we calculate the extra Power that's been consumed by CW Pumps?

Further, I want to get knowledge of designing of Aux. Cooling Water system. If any material is available about the design of Pumps and Piping system, please provide me the same

Details of your query are summarised as below :

There are two aspects in your query -

(i) Operation of the pumps : Basically the point of operation is the point of intersection between the H-Q curve of the pump (or the combined curve when pumps are running in parallel) and the H-Q curve of the system. The only information available about the H-Q curve of the system is that total flow required is (5810 + 675:) 6485 m3/h. Against this, from the curve of the CW pumps, corresponding to the differential pressures, the total flow would be expected to be 9760 m3/h.

Since this is excessive, you are throttling the valves. It is however not clear whether after throttling, the flow is really 6485 m3/h or not. To assess this, one should have either in-line flow measurement or one should, know the friction characteristics of the piping and of the valves in the throttled condition and the static head set-up of the system.

Another indirect way is to observe whether the temperatures after the heat exchange are as per the process-specification or not. If these are okay, the primary objective of the CW and ACW systems is satisfied.

The next point then is to see whether the objective can be satisfied in an energy-efficient manner. And satisfying the objective by throttling the valves is of course not the energy-efficient way.

The logical way is to make proper assessment of the requirements of the system and matching the pump selection or pump operation to the system requirements.

(ii) Design of ACW system.

I guess, your anxiety to look into the design of the system stems from the doubt about the mismatch between the pump and the system. However you have already identified that the total flow required is 6485 m3/h. Against that the nameplate details of each pump are 3450 m3/h at 20m head. So, if the total head is 20m, the two pumps together should be giving 6900 m3/h. This is just 6% extra margin over the required flow of 6485 m3/h. This is not a very bad situation.

The problem obviously then is because the head specified for the pumps was 20m. Actual head is much less, 14m (equivalent of differential pressure of 1.402 kg/cm2). And at such 70% head, the pumps are giving much higher discharge, 4880 m3/h (combined flow 9760 m3/h), one and half times of the required flow of 6485 m3/h.

This is typically an eminent case of how margins get added in the estimation of system head. And finally the system turns out to be one demanding energy-inefficient operation.

So it seems that you do not have to get into the design of the Cw or ACW system. What is needed is to set the running of the pumps to give 6485 m3/h at 14m head. When saying this, I would also like it to be checked whether pressure gauge readings are with valves throttled or open. If the differential pressure is 14m even with throttling of the valves, without throttling, the pressure differential would be still less!

The most forthright answer seems to be to retrofit speed reducers for the pumps. and if the, system demand of 6485 m3/h is a constant demand, the speed reducers need not be variable speed drives. They can be fixed ratio speed reducers of the mechanical types (pulleys, gears, etc.)

The ratio of speed reduction can be calculated. But better to confirm the condition of the differential pressure, whether it is with throttling of the valves or without.

.you would find a very, very innovative approach to designing of cooling water systems.

It is just an example how system can be designed for energy optimisation. Say there are 3 nos of coolers and a pump is delivering cooling water from cooling tower/other heat exchanger to cool those coolers. In system#1 the pump is delivering 20 ton/hour to cooler #1, 15ton/hour to cooler #2 and 10ton/hour to cooler no#3. As it is a parallel system total cooling water flow requirement is 20+15+10 = 45tons.

In system #2 the pump is delivering 30 ton/hour to cooler#1, as the flow a bit more than the previous system(20ton/hr) the outlet water remains little bit cool w.r.t previous system, and the cooler#1 outlet water can be introduced in cooler #2.5ton/hour water can be short-circuited assuming 25ton/hour will be sufficient for cooler#2 (which is again more than the previous system#1 15ton/hr). Out let water of cooler#2 can be used in cooler#3 and out of 25ton/hr may be only 20ton will be sufficient for cooler#3. So total water requirement is 30ton/hour.

Hence in system#1. pump is delivering 45ton/hour but in system#2 pump is delivering only 30 ton/hour and pump capacity will reduce accordingly.

Although the total example is based on assumption, proper calculation based on some real system may reveal even better result.

Horizontal pump to replace
vertical pump

We, Thirumalai Chemicals Ltd, are manufacturers of petro chemical products

situated at Ranipet, Tamilnadu. We have ordered, one horizontal centrifugal pump to replace our existing pump, which is working for main distillation circulation, as per the following specifications.

Horizontal jacketed centrifugal pump

Head : 15 mtrs, capacity : 250 cu.m/hr

Op.temp. : 240 deg c

Service : crude pa, specific gravity : 1.1

System : vacuum (70 torr to 180 torr)

Vapour pressure: 0.5 kg/cm2 absolute

Suction size : 200 mm : discharge size : 150 mm

Efficiency : 85%

Bkw(water) : 12, bkw(liquid) : 13.8

NPSHr : 1.3 m

Pump speed : 960 rpm, motor rating : 18.5 kw

Sealing : Mech.seal

Moc : casing/st. box cover- cf8m(j)

: impeller / sleeve - cf8m

: shaft - ss 316

Our existing pump is vertical axial flow centrifugal pump & all other specification as above . Let us know that the new horizontal centrifugal pump will meet our requirements, if any technical clarifications in this specification and type please advise us to proceed,

The data for calculating NPSH available is inadequate. Broadly, since the system is under low vacuum of 70 Torr for PA of sp. gr = 1.1, that makes Absolute pressure in the suction vessel as 0.07*13.6/1.1 = 0.865 mLC. With vapour pressure of PA being 0.5 kg/cm2 absolute, i.e. 0.5*10.336/1.1 = 4.7 mLC and with NPSHr as 1.3m and margin over NPSHr as 1m you would need a positive head to compensate (0.865 - 4.7 - 1.3 = -5.3) To provide also for Frictional losses and velocity head for 250 m3/hr thru 200 mm pipe ((250/3600/(pi()/4*0.2*0.2))^2/2/9.81 = 0.25m, the pump will have to be installed at say minimum 6 m below the bottom of the suction vessel.

The less stringent vacuum of 180 Torr can save some 1.35m.

For 250m3/hr 15m head at 960 rpm the specific speed works out to 33.19, which is fairly on the border level of radial flow pump and 85% pump efficiency for this specific speed and flow rate seems quite good!!!!! Even NPSHr of 1.3 m for a pump of this specific speed and flow rate seems too good!!!!!

Even a vertical pump would need minimum submergence of 6m. I wonder how your existing vertical pump is installed.

1) Question - One centrifugal pump with 60mtr head is giving 7 Kg/sq.cm. in running whereas it is supposed to give max. 6Kg/sq.cm. So now the question is where from the 1 Kg/sq.cm is added? Please comment.

2) Question - For the centrifugal Pump, what should be the reason if the indicator of pressure gauge is fluctuating? Why not it's been steady ?

There are two possibilities -

(1) If the shut-off head of the pump is more than even 7 bar, in a given system situation, the operating point can be at 7 bar pressure.

(2) The total head of a pump is a differential head. If the pump is under positive suction of say, I bar gauge (i.e. 2 bar absolute) and the pump is developing additional differential of 6 bar, the pressure reading will be 7 bar.

I have a query, it is divided into three parts :

a) Is it advisable / mandatory to keep the by pass line of ARC valve, at Boiler feed water (BFW) pump discharge, facing/going down?

b) Would there be any change in the configuration, in case the pump discharge is at top or at side?

c) Are there any constructional restrictions in ARC valve, which forces the above mentioned configuration, especially in high pressure applications such as BFW?

Two basic concepts are to be kept in mind.

1. ARC valves are eminently uni-directional valves. (2) Flows happen from upstream to downstream, i.e. high hydraulic energy level to lower energy level.

3. Orientation by position "at top or at side" becomes significant based on the design pattern of the ARC valve. Being basically unidirectional, they are similar to non-return valves or pressure relief valves. They are actually pressure relief valves only. with the provision of the relief line being connectible to the circuit to make the flow to return to suction and not go waste. Pressure relief mechanism often being spring loaded the orientation may not matter. But it is better to follow the manuals, because the spring-loading may have difference in response based on orientation.

For metering pumps what are the reasons to use double check valves?

Taking clue from your mention of the double check valve, I browsed the net 'serach'ing for "DOUBLE CHECK VALVE" and got the following description of a dual check valve. "Dual Check Valve Baackflow Preventer is designed to prevent cross-connections of non-potable water (non-hazardous) into the safe drinking water systems. It is a compact and economical device that is easily installed, serviced and repaired in the line. The device consists of two independently acting, spring-loaded check valves in a corrosion resistant material."

From this it seems that one uses a dual or double check valve, when there would be apprehension of unwarranted mixing of non-compatible products. I would hence guess that one would not need a dual check valve, if there is no such apprehension.

Design of Submersible Motor-Thrust bearings

I have tech question regarding the thrust and sleeve bearings of the deep well submersible motor, some 6 year or before every European and US made motor had bronze based thrust bearing , they were more rigid to vibration and rough transportation, well now most of the manufacturer like , Franklin , Tesla, jet spa , etc switched to carbon based bearings now what i want to know, is either these carbon based bearings are more though and good compare to old bronze based bearings or these changes are made due to commercial point of view , pl. let me know what is the idea behind this,

In terms of properties desired of a bearing material, carbon has lower co-efficient of friction, better wear-resistance, better corrosion-resistance with most liquids, especially bore well waters and higher temperature-withstanding capacity or better creep strengths. So the change-over to carbon is technically sound.

We are the manufacturer of submersible pumps and open well submersible pump for the last 5 yrs. We have our own R&D set up for improving and upgrading our product quality. During our R&D we are not able to get the exact technical data of some critical functional components. ,So we are forwarding our queries in our mind to the supreme like you. As below :

1 What is.the function of diaphragm & how it should be arranged ? Which type of material in more suitable for sandy condition.

2. In submersible pumps and open well submersible both side thrust bearings are required or not ? How to calculate the minimum surface area required for bearing ? Which type of bearing will improve starting torque (fixed/loss segment) how?

3. What are the books available for hydraulic designs of submersible pumps and centrifugal pumps ?

Bore well Submersible or open well submersible pumps are basically impeller pumps and hence, centrifugal pumps. So for hydraulic design one can follow any book on design of centrifugal pumps. Some commonly used books are -

1. Design of Pumps fans and blowers by Austin Church & Jagdish Lal

2. Centrifugal and axial flow pumps by Stepanoff

3. Impeller pumps by Troskolansky and Lazarktewitz - this book has been out of print for many years. Some institutions may have it.

4. Pump Handbook by Karassik

5. Book by Lebanoff

6. Book by Anderson - outlining area ratio theory

7. A Russian book by Mir Publishers, "Hydraulics" by Nekrasov deals with design of pumps quite crisply.

8. These days people also use CFD (Computational Fluid Dynamics) packages

9. It seems the credit of first ever documenting the theory and design of centrifugal pumps should go to Pfleiderer for his book in German "Die Kreiselpumpen"

Diaphragm in water-filled submersible motor serves the function of providing extra volume as needed by expansion of water due to rise in temperature. If increase in volume is not provided, the pressure would increase and will cause leakage of water, depriving both the journal bearings and the thrust bearing of the tribological effect. The diaphragm should be arranged so that the volume inside can expand. Exterior of the diaphragm is exposed to the bore well water. Apart from wear due to sandy water, the corrosive effect of bore well water can range from salty (alkaline) to sulphurous (acidic and exothermic). It seems Neoprene rubber can be an adequately general purpose material.

Design of the thrust bearing will be primarily dictated by the energy of mechanical friction getting converted into thermal energy, in turn affecting both efficiency and wear life of the motor. Frictional force F - Mue*Thrust and power loss in friction will be F*s, where s-distance traversed in unit time, say second. With mean diameter of thrust bearing being d, distance traversed in one second will be pye*d*n/60. Mue can be minimized both by best combination of materials of rotating and stationary faces and by having the rotating face experience rolling friction instead of journal friction.

We are manufacturing Agriculture Centrifugal pumps. in the manufacturing of Centrifugal pump we generally put A Gun metal Bush in the body. now we have started to put a mechanical seal before the bush in the body of the pump. (impeller side.) and we reduced the size of bush to that extent. Now my inquiry is :-

1. Apart from saving the bush from sand which comes with water, what is the main use of putting the Mechanical Seal in the pump.

2. Does there is any disadvantage of reducing the size of bush and does there is any method of determining the right thickness and length of bush.

3. We make pumps from Cast Iron. can we use plastic impeller in the pump made from cast iron.

Your description does not bring out the construction of pump clear. However, purpose of mechanical seal is not to protect bush from sand. It is actually the other way round, i.e. the bush, (commonly placed between the impeller and the mechanical seal) is called as the throat bush. Because the shaft runs close with the bush it has very little clearance for sand to pass across. The bush also cuts down the pressure of leakage from behind the impeller into the sealing area. By that it brings down the load on the mechanical seal. To be able to do that as much effectively as possible, the bush should have as much length as possible. The common sequence of position of components is thus the impeller, then the bush behind the impeller and then the mechanical seal behind the bush. Please check up whether your construction is also similar.

Impeller for sand contents

In my area iron & fine sand contents are found maximum in bore wells. Which type of impellers in pumps are durable?

One way to contain abrasive wear is to use slow-speed pumps. In coal mines they have been using helical rotor progressive cavity (HRPC) pumps as face pumps. The liquid is very much similar to what you find in the borewells there.

Wear due to sand and coal fines is different from wear due to slurries like tooth paste or ash slurry. Sand particles would impinge on to the surface with a steep angle of incidence, whereas slurries will abrade the surface moving very close and parallel to the surface. Wear due to impingement can be better tackled by resilient materials like rubber. Conversely abrasion needs hard surfaces. Effect of impingement will be less severe at slow speeds.

HRPC pumps inherently have an elastomer-lined stator. They are also positive displacement pumps and hence can pump from any depth. They are inherently slow speed machines. HRPC pumps have been made in India to inject sea water into oil wells to improve the productivity of the wells. And the depths reached are 1000m 1 km below sea level. HRPC pumps would not need submersible motor. The pump shaft inherently has a double universal joint. So they would be easier installed even in bores with some misalignment. It seems HRPC pumps become the best candidates for difficult bores.

I am manufacturing water cooled openwell Monoblock in single phase. My core length 45mm, can u give me the winding design?

Alongside of core length, information is required about power rating rpm, stator stamping OD, ID and slot details, i.e. number of slots and shape and area of each slot.

Monoblock Vs Openwell Submersible Pump

In agriculture for openwell still farmers are using Non-Self Priming Centrifugal Monoblock Pumpset and always changing the position of Pumpset whenever water goes below the level. I personally feel that they are wasting lot of energy and they should use Mono Submersible Pumpsets.

I had discussions with 3-4 dealers of Monoblock Pumpsets and as per them all NON Self Priming Pumpsets are being design for negative Suction offcourse with suitable foot valve. My explanation to them was yes you can use Monoblock Pumpset with foot valve but when submersible Monoblock Pumpset are available of better efficiency you should not sell/use these pumps for open well.

As per my views people who are not aware for Open well submersible Pumpsets ore still using Non Self priming Mono Block Pumpsets.

To update my knowledge please explain

(a) why stilt people are using non-self priming Monoblock Pumpsets in comparison to openwell submersible Monoblock Pumpset atleast for agriculture use

Because of :

1. Poor Knowledge

2. Lesser price

3. Still they prefer to use old design of pumps

(B) If neither cost nor any limitation, will it be recommended to use non-self priming Monoblock Pumpsets with negative suction.

(c) For testing of total head can be Monoblock Pumpset can be tested with positive head say 1 Mtr. (Total Head = Diff. head - Suction Head)

or it has to be tested with Negative suction or both methods are O.K.

The 'motor efficiency factors' for open well submersible pumps as per IS-14220 are less than corresponding factors for surface monoblocs as per IS-9079. So, on the count of overall efficiency the surface monoblocs as per IS-9079 are as of now more efficient than open well submersibles as per IS-14220. The advantages with openwell submersibles are freedom from priming troubles and freedom from anxieties of ensuring suction lift not being excessive or from flash flooding of pump and motor. For people, who have been used to using surface monoblocs since years, one need not over-emphasize switching over to open well submersibles, primarily because it does not serve National concern for Energy Conservation in agricultural pumping.

If suction lift is not so excessive as to cause the pump to suffer cavitation, testing a pump with suction lift or with positive head would not make any difference to the pump's performance characteristics.

Books on Pumps

I want to know whether any design books are available for regenerative pumps. Especially self priming pumps like mini monoblocks.

I haven't known of any books detailing the design of regenerative pumps. They are good candidates for research on how to improve the efficiency of these pumps. But I guess the efficiency will remain poor, because for centrifugal pumps even approaches in CFD try to minimise internal recirculation in the impeller passages. as against this, working principle of regenerative pump is based on recirculation!!!!!!!!!! A "regenerative" pump cannot "regenerate" without recirculation!!!!!!!!

Even when I say that, I remember very clearly that years back, sometime in 65-66, Late Mr. S. G. Phatak then with Kirloskar Brothers had contributed an article in Indian Pumps outlining Navier Stokes equations for regenerative pumps. I doubt if anybody has a copy of that article.

I am indulging in making too many contradictory statements, because I also know that source codes of most CFD packages are based on Navier Stokes equations. Mr. Phatak dealt with those very equations for regenerative pumps, when computers were not even invented!!!!!!!

Yet approaches in present CFD packages are contrary to the working principle of regenerative pump!!!!!!! Only one thing which is in favour of regenerative pumps for CFD analysis is that space between two vanes makes a very, very sharply defined boundary condition. But the antithesis of this is that the flows would be very very turbulent. So, I guess, one would not be able to read any streak lines in a CFD output.

Please suggest the product for checking system head (on line) i.e. pressure gauge and digital gauge as well as their suppliers.

Pump develops differential head. So difference between readings of two gauges - one on suction line and another on discharge line, referred to a common datum, often the pump centreline, would give the differential head. One can as well use a differential gauge.

Tech Quest

Q. I am working in ESSAR OIL, working as a maintenance engineer in CDU/VDU units in Refinery located at Vadinar.

I would like to know the following:

1. After vaporization of any lube oil which we are using in over hung C.F. pumps in our refineries, how much % reduction in viscosity of lube oil and what wilt be the adverse affect of presence of condensed drops in bearing housing?

2. what are parameters we are checking while doing oil analysis? And what are the problems upcoming in any equipment either pump or compressor?

3. How much % of tife of any equipment will be increased by predicting and attending of upcoming problems indicated by the result of oil analysis

1.) After vaporization of any lube oil which we are using in over hung C.F. pumps in our rifineries, how mach % reduction in viscosity of lube oil

Different grades of lub oil have different temperature v/s viscosity relationship.

Query 1a and what will be the adverse affect of presence of condensed drops in bearing housing?

Presence of condensation not only affects the lubricity of oil, condensation will also cause corrosion of shaft.

2.)what are parameters we are checking white doing oil analysis?

Lube oils are to be checked for viscosity and ingress of foreign matter into the oil.

Query 2a And what are the problems upcoming in our any equipment either pump or compressor?

Lack of lubrication will cause any equipment to suffer high friction, rise in temperature, damage or breakage to bearings and shafts

3.) How much % of life of any equipment will be increased by predicting and attending of upcoming problems indicated by the result of oil analysis.

Proper lubrication is only one of the causes of failure of a pump. Even if the lubrication is proper, a pump may fail due to other causes, such undue vibration, cavitation, misalignment, leakage at shaft-seal, improper operation, especially, operation at low flow, etc. philosophy of predictive maintenance is to monitor condition of the pump for all these critical parameters and achieve lengthening of Mean Time Between Failures (MTBF

Q. I have a query regarding the leakage current test in submersible motors (3 phase).

There are two methods i suppose: -1st :- By using a mille ampere meter. Connect one wire of mille ampere meter to motor body and the other wire to earth. Now give rated supply to motor (no load). Take the reading of mille ampere meter which is supposed to be leakage current of that particular submersible motor.

2nd :- By using the commercially available leakage current meters. Connect one wire of this leakage current meter to one of the phases i.e R,Y or B and the other wire to earth. The reading of leakage current meter is taken as the value of leakage current. Now the question is which of these methods is correct?

I performed the leakage current test at my factory premises by using both of the above methods. The values I got were not same. People at ERDA, Vadodara are using the 2nd method. But the being an electrical engineer doesn't find anything wrong with the first method, which appears to be logically correct to me.

I too would feel that method 1 is more appropriate. Leakage current, to my mind, is current which is escaping from such point on the motor, from where current is not supposed to theoretically leak. Appropriate point then is the motor body.

Worry about the leakage current is also a consideration of safety. Leakage of current on motor body ought to be within safe limit. From this view-point also motor body appeals to be more appropriate point. I am replying from logical considerations. One needs to check up, whether standards do specify any exact method and location. If the standards do not specify clearly, then logic should prevail.

Q. Could you please give some info on the following, For a closed a loop chilled water system, if the elected pump's head is more than the required (Say selected pump's head is 4.5 bar and actual head requirement is only 3,5 bar) what are the possible damages to the system ?

If at desired flow, Q, system needs 3.5 bar and pump curve shows 4.5 bar, the pump will operate at much higher flow and would possibly overload the motor.

If the pump has to feed a chemical, say an alkali to neutralise an acid, excess flow will make the chemical reaction to yield an alkaline end-product, instead of neutralisation.

To set the system at desired flow, one can either throttle the valve or trim the impeller or reduce the speed.

Throttling the valve is not efficient way of doing the setting. Also, on throttling the valve, the pressure seen on the gauge is before the valve and not after the valve, If one puts a gauge after the valve also, one would notice that, that gauge would not show any alarming pressure.

In most applications, one is more concerned of the flow desired, than the pressure. And one should be also concerned f or overloading of the motor. Options tor getting desired flow and containing overloading are the three options mentioned above.

Q. I am your regular subscriber since long. I have a query about selection of pumps . kindly guide me the best pump according to my requirement-

We have to suck clean water from river like open water canal. Currently we have installed Lubi make 30 HP V10 submersible of 1 stage 6" delivery giving about 4500-5000 lpm. Now we want to installed any big pump that gives the about 15000 lpm.

Horse power- no bound may be up to 50 hp Source- Suck Open clean water from river
Height- 17-20 mtr
Water requirement- 15000-20000 lpm
Type- Submersible or any electric pump

Data is not quite clear, especially about static head. An accurate estimation of total head should be done before selecting a pump. Since the pumping is from a canal, the level on suction side may not be varying much. Also level to which water is to be raised also may be a fixed level. The distance to which water is to be carried needs to be taken into account. Different sizes of delivery pipe would give different frictional head.

Efficiency of the motor should also be a point of consideration. Basically a dry weather proof motor may be more efficient than a submersible motor, Alternatively, a dry submersible motor, as in sewage submersible pumps can also have efficiency competitive with weather proof motor.

Assuming levels will remain same and in turn total head will remain same, for 3 times flow (15,000 lpm in place of 4500 to 5000 lpm at present) would need 3 times motor hp, hence nearly 90 to 100 hp. So 50 hp would not be adequate.

There can be an option of using more than one pumps. A large pump of 15,000 lpm would anyway be a new larger installation. Instead, one can consider modular enhancement of existing installation. Maybe a close-coupled vertical pump with dry submersible motor or a sump type or vertical turbine type pump with weatherproof motor could be also considered. If a pump-room or jack well can be constructed and if the suction lift is about 4 or 5 meters, even a horizontal split casing pump with normal horizontal motor could be considered.

The situation seems to be a good example for exploring number of options and making Life Cycle Costing of various options, before making final selection.

Q. For vertical Centrifugal Pumps (vessel mounted), is writing "Flooded suction" a correct statement or actual estimated NPSHA should be indicated?

It is always good practice to mention NPSHa. Some pumps need minimum submergence. Mere fact of suction being flooded, may not yet provide "minimum submergence". Inadequate submergence can become cause for cavitation.

For volatile liquids of high vapour pressure also, just flooded suction would be inadequate NPSHa. For example, for pumps for pumping LPG or for pumps for condensate extraction, purchase specifications would often declare NPSHa = zero, implying that the pump design should be taking care of NPSHr and required margin in the design itself. The design of pumps for such applications is often made "encastre" type.

Q. What is bearing crush? how do you check bearing crush?

(Additional Information : This question was replied in Pumps India, Jan/Feb 07 issue, however we have received more information on it from Mr Rai, we are thankful to him for his contribution).

Regarding bearing crush. In this context I would like to add that this term is used for the interference of the leeve type bearings either with sphericai seating or ylindrical seating. Out side diameter of the sleeve type bearing inserts are slightly higher than the inside diameter of the bearing housing. To achieve this the diameter of the two halves at the position perpendicular to the parting plane is higher than the diameter at the position across the parting plane. This difference is termed as Bearing Crush. This serves in the following ways.

It gives tight fit of the inserts into the bearing thus preventing any movement of the bearing insert. It prevents lube oil to be filled in-between the insert and bearing housing thus better heat transfer from the bearing. To avoid the oil clearance being decreased it is of utmost importance to take oil clearance measurement after assembly.

It is written to share the information and should not be taken otherwise.

Q. We are one of the leading blood bank equipment manufactures in India. We are using a pump of 28watts,230v,50Hz single phase AC for one of our equipment. but we are frequently getting the complaints from the field especially leaking through the shaft sealants, Can you suggest me a suitable pump for our use.

28W, 230V, 50 Hz, single-phase becomes more the specification of the motor than the pump. For the pump, the specification should mention flow-rate and pressure, and also the liquid characteristics, obviously blood. Blood is a highly viscous liquid and also needs to be handled hygienically. I think a pump to handle blood hygienically should be a positive displacement pump preferably of the diaphragm type or peristaltic type, which are inherently seal-less and leak-less. This is of course a general recommendation. This can be better fine-tuned, if you would detail the flow-rate and pressure.

Q. Thank u for the valuable information the requirement is not for the use in circulation of blood. we required it for circulation of water at a temperature of 56 degree in one of our equipment called Plasma thawing bath. we want the pump spec of head 4.5M and LPH -300.

Since you had earlier mentioned about the problem of leakage with these pumps for Plasma Thawing Bath, I would recommend a pump with vertical motor. Such pump is used in circulating machine tool coolant in machine tools, such as lathes, drilling machines or CNC lathes, etc. The height to be delivered is about same as 4.5 m mentioned by you and the discharge is also of similar order as 300 lph. Pumps on machine tools also handle circulating coolant, which often could be hot after contact with metal-chips machined by the machine. As such 56 Deg C is not a very hot temperature for most common materials used in pump-construction.

Other similar pump is one, used in Desert coolers. For pumps for desert coolers, there has even been an Indian Standard IS-11951, where the pumps lifts the water from the tray in the bottom to the tray on the top, so that people can get cool breeze from the air blown by the fan across a curtain of the fragrant grass.

Pumps of machine tool coolant and of desert cooler are made by many, many people. For your application, to be free of leakage problem, they should be vertical, whereby the pump stays submerged in the hot water sump.

Hope, this helps.

Q. We Manufacture Domestic selfpriming and Centrifugal pump during the test of centrifugal pump of 1.5 hp 2 x 2 The following is the observation. Impeller OD 125 mm Pump over heated in 240 v and tripped of after running for six hour Reduced OD to II8 mm 240 v pumps is ok 220 v it tripped off after an hour Impeller OD reduced to 117 mm 240 v and 220 it is ok but tripped off at 180v we do not want to apply this method by reducing the OD of the impeller. What is the reason for the tripping off Was it required to reduce the impeller od by doing so will the performance of the pump get affected. Please advise with the solution.

Tripping off of the motor is due to overload. With regenerative turbine type impellers in domestic, self-priming centrifugal pumps, it may prove effective to try reducing the impeller width by facing from both sides and redoing the assembly maintaining the original clearances.

Q. Thanks for the reply sir, can this method applied for centrifugal agri monobloc pump, by doing facing on both the sides, Will the centre point of the vanes with the volute can be maintained pls clarify.

No. The design logic of commonplace centrifugal pumps is totally different. By the way, have you already tried facing the impeller of regenerative turbine type impeller? What percentage change in width has given what change in power? I shall be curious to learn of the trials.

Q. One of our customer using centrifugal (end suction)pump to pumping the hexane from underground tank, they have facing cavitation problem. Pipe line arranged on the top of tank. Pls, advise how to avoid the cavitation

Most effective solution would be to change over to a vertical sump pump, which will work submerged in the liquid. The pump should stay submerged even with the minimum liquid level. Such pump will eliminate suction piping, footvalve, cavitation and re-priming at every loss of prime.

Q. Kindly tell me the formula for horizontal piston pumps, available in single, double and triple pistons. how can v calculate the discharge at the pump at given pressure. for eg. a pump with suction capacity of 36 lpm and max Rpm 950. how can v calculate the discharge of pump at diffirent pressures by using their pulley sizes etc.

Discharge of a piston pump is basically (area of piston * stroke length * rpm)

Maximum Pressure = HP*k*Pump efficiency/Q/(sp. gr.)

k = constant appropriate to the units used for power and Q

For pump efficiency one shall have to consult the pump-manufacturer.

For multiplex pumps, total Q=Q per piston*no. of pistons, if their strokes are in phase.

Mostly they will NOT be in phase. Then the calculation becomes complex. Ideally one may have to make "discharge versus time" plots for each piston and derive a summation plot, keeping in mind the phase differences.

Tech Quest

API 610 recommends Minimum size requirement for casing connections other than suction & discharge nozzles based on Discharge nozzle size excluding seal flushing & gauge conections. Why?

What I appreciate most about API Standards, is that these standards are evolved by an organisation of pump-users, with little influence of pump-manufacturers.

Size of casing drain connection is important to ensure that the casing can be drained as fast as possible, saving downtime during maintenance. Size of pressure-gauge connection is important because that automatically dictates dial-size of the gauges and in turn their accuracy and readability.

Once the significance of these sizes is also appreciated, one understands and appreciates the standard also better

What role would moment of inertia of the motor play in damaging, rather, shearing the power screw of a triple screw pump? Salient details are -

1. Pumped liquid - lubricating oil

2. Capacity 160 lpm

3. Rated pressure - 180 bar

4. Drive - 60 kW, 2900 rpm

5. Mounting - Vertically into a 36 m3 oil tank. Minimum submergence recommended to avoid vortex-formation, air entrainment, dry running is 40 mm. The recommendation is not always followed.

6. Constructional features - There is no packed gland or mechanical seal. Pump shaft or the power screw has a balance piston on the driver side, with close clearance. between balance piston and pump casing. Leakage if any, maximum 0.5 lpm, would flow over the pump and drip back into the oil tank.

7. Material strength and diameter of power screw -24mm, 16CrMnS5, with minimum UTS of 570 MPa

To get a clarity for myself, on the role of moment of inertia of the motor, I would paraphrase the question a little differently. "Though a pump requires only 60 kW rnotor, would a 100 kW motor, inadvertently or even wantonly connected, cause damage, rather shearing of the pump shaft?" My logical answer would be "No".

A 100 kW motor would have much higher moment of inertia than a 60 kW motor. If the shaft has suffered seizure, a 100 kW motor has rather, a better chance of overcoming the seizure and make it run than a 60 kW motor.

The root cause of the failure seems to be the seizure of the shaft than excessive moment or torque imparted by the motor.

Standards for motors specify pull-up or starting or locked rotor or breakaway torque to be 150 percent of full-load torque. This testing is typically done by locking the rotor. This actually simulates a seized shaft and demonstrates capacity of the motor to overcome seizure. A motor does try to overcome a seizure. Excessive torque may get imparted in this effort of the motor to overcome the seizure. But the demand for excessive torque comes from a seized shaft. If the shaft is not seized and does not demand excessive torque, the motor will not impart, by its volition, any torque more than what the driven shaft demands. If the driven shaft demands only 10 kw load from a 60 kW motor, the motor would provide just as much. That is what is called as part-load running of the motor. Efficiency of the motor would of course be poor in such part-load running. Motors have power demand of their own, even when running on no load or zero load. So, motors can run all the way from zero load or no load to full load and somewhat beyond full load, which is overload. Basic fact is that motors respond to the demand. They do not impose load on to the driven equipment. They impose load on to the supply system, not on the driven system.

Root causes of the shear of a driven shaft would be misalignment, thermal load, seizure. Possible causes for a shaft of a screw pump to suffer seizure would be dry running.

Incidentally, if both the power screw and the idler screw have same metallurgy and area of course, running in frictional contact, they are susceptible to suffer electrolytic galling and consequently a seizure, more so in dry running.

To prevent dry running in the given installation, it seems that a level controlled interface with the motor's starter would be a good protection. The other check should be on using dissimilar metallurgies to avoid electrolytic galling at surfaces in frictional contact.

Bearings in Pumps

We are a well established company in the production and distribution of various kind of Ball & Roller bearings. Would highly appreciate if could let us know which bearings are used in Pumps production. What is their application and quantity per Pump. We await your reply per return.

Bearings used in pumps are of various types. To list -

1. Monobloc pumps would have no separate pump-bearings, because the pump assembly is on the extended shaft of the motor. So, bearings of motor serve a also as pump-bearings.

2. Pumps coupled to the driver through a coupling will have pump bearings. Number of bearings varies depending upon application.

For economical pumps as for agricultural purposes, the pump may have only one anti-friction bearing, the hypothesis being that the throat bush in the stuffing box also acts as a bearing support.

Industrial pumps which have to often run round-the-clock would have a distinct bearing housing with 2 bearings, one at the driving end, i.e. near to the driver and the other nearer to the pump. If axial thrust in the pump is estimated to be significant, e.g. with semi-open impellers, the pump would have an anti-friction thrust bearing, often of the angular contact type and in matched pair.

Large vertical turbine pumps would have tilting pad (Mitchell or Kingsbury type) thrust bearings.

For very large horizontal pumps where anti-friction bearings of large shaft dia are not available from regular product ranges, people may use journals with splash ring etc. Axially split casing type pumps and multi-stage pumps are also called as "between bearings" pumps, meaning the pumps would have two bearings at the two ends.

For bore well submersible pumps with water-filled wet motors, the pump assembly has mainly stage bushes and the motor also has bush bearings, because the motor is filled with water. Oil-filled motors would have anti-friction bearings.

Helical rotor progressive cavity pumps would often have only one bearing, because the pump shaft needs to drive the rotor through a universal joint.

Twin screw pumps may have as many as four bearings, two on each shaft/screw.

A triple screw pump however may have only two bearings, since there is only one driving screw and other screws run as idlers.

So a variety of logic for number of bearings and types of bearings in a pump.

I want to know the steps taken in a pumping system to attain the desired operating point for the system when

(a) the pump is driven by a fixed speed motor and,

(b) when the pump is driven by either a variable speed motor or a turbine.

The operating point is the point of intersection between the H-Q curve of the pump with the H-Q curve of the system. Once a pump is set into a system, this will happen automatically. But if the operating point, which happens automatically is not the 'desired' operating point, one has to modify either the pump curve or the system curve.

There are two ways to modify the pump curve -

1) Change the speed of the pump

2) Change the diameter of the impeller of the pump

3) The system curve can be notified by modifying the system. This is usually done either by changing the setting of the delivery valve or one can change it also by revamping the system by changing the pipe-sizes and/or layout of the piping.

4) If the suction conditions in the system are prone to cause the pump to cavitate, modifying the system to eliminate cavitation will also modify the pump curve from a cavitating condition to non-cavitating condition.

5) For changing the speed of the pump (option 1 above), changing the driver from an electric motor to a turbine will often become changing from a low-speed driver to high-speed driver.

Such change is possible even by using a gearing or pulley mechanism between the pump and the motor. But at increased speed the pump demands higher power input. So, it becomes important to check whether the motor has adequate margin in power. No such caution is needed if "desired" operating point is obtainable by reducing the speed.

For determining the required speed at the "desired" operating point, say

(Q",H")one needs to find the point (Qo, Ho) on the pump curve H = a*Q^2+b*Q+c which also is a point on the parabola through the origin and (Q", H"). The equation of this parabola will be H = k*Q 2, where k = H"/(Q")^2.


Since (Qo, Ho) is to be a point both on

H = k*Q^2 and H, a*Q^2+b*Q+c

to find (Qo, Ho) one needs to solve the quadratic (a-k)*(Qo)^2+b*Qo+c = 0

Actually all the mathematics starts with knowing the values of the co-efficients a, b, c for the pump curve H = a*Q^2+b*Q+c This is not difficult, if one knows three points on the curve, say,(0, Hso), (Q1, H1) and (Q2,H2) and solves simultaneous equations. A simpler way to do this is to plot the pump curve in an Excel spreadsheet and fit a 'trendline', setting also the option for the display of the equation of the polymonial of degree 2.

MOC for Abrasion & Corrosion

I would like to know the diffirence between abrasion and corrosion. What type of M.O.C for impeller and shaft is suitable for abrasion and corrosion?

Also let me know the selection parameters for Impeller and shaft M.O.C.

One commonplace example of understanding the difference is water laden with sand. Sand, per se, is not corrosive, but it is very abrasive. Conversely acid with no entrained solids, clear acid, will not be abrasive but highly corrosive. Sea water will also be corrosive. But corrosion due to sea water is due to its alkalinity whereas corrosiveness of acids is acidic in nature. MOC for corrosion resistance has to take into consideration whether the corrosiveness is acidic or alkaline.

Abrasion is also of two types. Abrasion due to fly ash in power stations will be from fine particles moving too close to the surfaces and abrading the surfaces. Abrasion due to sand particles or coal particles will be due to the particles hitting hard on the surface and bouncing back and hitting repeatedly. This is rather erosion than abrasion. So nature of abrasive wear depends upon the angle of incidence of the particles w.r.t. the surface. Usually hard surfaces would take abrasive wear better and resilient surfaces such as elastomer-linings would take the erosive wear better. But this is too much of a thumb rule. One needs to study the wear patterns and refer to the data available in handbooks.

How to determine Minimum stable continuous flow & Minimum thermal continuous flow. Is there any standard which explains about these parameters & the determination in detail

Minimum stable continuous flow is to be read on such H-Q curve which is unstable. In unstable characteristics, Hmax is greater than Hso (Head at shut-off). In such case, Minimum stable continuous flow will be where a horizontal thru' Hso will intersect the H-Q curve of the pump.

Minimal thermal continuous flow is that flow, when the liquid will experience churning caused by internal re-circulation. This happens because, the cross-sections of the hydraulic passages prove to be too large for the amount of flow to be carried. The designer designs the passages ideally for the design flow. At flows less than the dsign flow, the passages are not ideal. This is also one reason for the drop in efficiency at flows different from design flow. The effect becomes accentuated at flows less than Minimum thermal continuous flow. Churning of the liquid causes temperature of the liquid to also rise. This in turn raises the vapour pressure of the liquid. In turn the available NPSH gets affected. By all these considerations the curve for NPSHr v/s a becomes uncertain. So, manufacturers show NPSHr curve only ahead from Minimum thermal continuous flow.

Obviously both Minimum stable continuous flow and Minimum thermal continuous flow are to be recommended by the manufacturer and cannot be obtained from standards.

In API-610 one finds a mentibon of continuously rising characteristics, that means a stable characteristics, i.e. where Hmax is only at shut-off. To be more mathematically correct, for a stable characteristics, the point of maxima is not in the first quadrant.

Tech Quest

1.     Two motors are having same H.P. but in different RPM, Which will take     more amps?

2.     what is the use of lantern ring in the stuffing box?

1. There is no direct relationship between motor HP, motor RPM and current.

Wattage = 1.732*V*I*cos(phi) is a basic formula.

And 746 W = 1 BHP. Metric HP= 736W.

Values of V and I for putting in the formula will be values related to input to motor. But motors are rated by their output HP. So, if V and I at input to motor will be proportional to (Motor HP/Motor Efficiency). If two motors are of same HP, one 4 pole and another 2 pole, have same efficiency and are working with same voltage and power factor they should draw equal current (amps).

For same HP, different RPM will give different torque from the motor. But that is on the output side of the motor. Current is measured at the input to the motor. That is why there would be no direct relationship between motor HP, motor RPM and current.

2. Lantern ring acts as a container for the flushing, cooling, lubricating feed. Since lantern is fed with liquid under pressure, it also acts as the barrier against ingress of atmospheric pressure, thus safeguarding the vacuum being developed at the eye of the impeller. Since liquid at the lantern ring should also serve flushing of the particles loosened from the rope packing, a drop by drop leakage is recommended for stuffing boxes packed with rope packing.

Injectable sealants which are coming into vogue do not need lantern ring, because no offensive particles get loosened out, they have lubricant impregnated and work with much less co-efficient of friction and hence do not need much cooling. In effect they work with almost zero leakage.

Q.    Why does a star delta starter not function if the delivery valve is slightly     open while starting the pump. When shifting from star to delta the breaker     trips. Assuming That the delivery valve is open to 1/10 th of its capacity     that will not exert a great amount of load on the pump but the starter     ceases to function unless the valve is closed tightly.

The starter would trip, if it is not able to exert enough torque to drive the load. Although delivery valve being open only to 10% of its capacity does not appear to be excessive load, prima facie, the 'sense and respond' logic of the starter itself is causing it to trip. Maybe, the excessive load is caused due to bent shaft, misalignment at the coupling, worn bearings, rubbing at wearing rings. All other causes for overload apart from amount of opening of delivery valve need to be explored.

Pump Efficiency

Q.    How to improve Pump Efficiency? In all pumps. Please explain briefly. I     want to know how can I identify the H.P of Motor in the field.

There would be two points of view regarding improving efficiency of a pump, one, in the field and the other, at manufacturing.

Improving efficiency in the field is the same as using a pump efficiently, which in turn is same as energy conservation in pumping. A comprehensive note on this can be found at www.pumps.org There, seven ways for saving energy in pumping are detailed. Briefly they are as follows.

1.     Design systems with lower capacity and total head requirements. Do not assume these     requirements are fixed.

2.     Avoid allowing for excessive margin of error in capacity and/or total head.

3.     Despite the tendency to emphasize initial cost, you will save in the long run by selecting     the most efficient pump type and size at the onset.

4.     Use two or more smaller pumps instead of one larger pump so that excess pump capacity     can be turned off.

5.     Use variable-speed drives to avoid losses from throttle valves and bypass lines, except     when the system is designed with high static heads.

6.     Use pumps operating as turbines to recover pressure energy that would otherwise be     wasted.

7.     Maintain pumps and all system components in virtually new condition to avoid efficiency     loss.

Efficiency in manufacturing is through better design, better surface quality, especially of hydraulic surfaces of casing and impeller, close running clearances, least leakage, less wear at bearings, proper dynamic balancing and alignment.

Q.     Whether API-682 seals can be fitted into ISO-2858 pumps? If so, which     edition 1 or 2?

ISO-2858 is adopted in IS-13518. Standard on "Dimensions of Stuffing Box Cavities" is IS-11382. It suits pumps as per IS-13518, i.e. ISO-2858. ISO-2858 was also derived substantially from DIN-24256 and the cavity dimensions derived from DIN-24960. DIN-24960 can be considered as the 'father' standard for API-682. But Americans are still attached to FPS system of units, whereas in European and in ISO standards the approach is emphatically metric and wherever possible, even SI system of units. American also have started giving metric converted values in their standards. But basic or reference dimensions are in FPS units and metric values are only conversions. So, dimensions in IS-11382 derived from an ISO-standard (I do not have the reference ready at hand), itself derived from DIN-24960 is metric in its spirit. So there would not be exact matching. Typical example is of Nominal size 3", which becomes 80 mm nominal in metric, but the difference is large, 3" being 76.4 mm. Even 25 mm being metric nominal of 1" also gives a difference of 400 microns, which is a large difference for shaft diameter, since there has to be much fi Further to my reply I would like to add the following. For requirements of pumps not necessarily as per API-610, ISO-2858 seems appropriate. ISO-2858 is already adopted in IS-13518. Actually, conditions such as vibration limits, minimum nozzle loads and moments as specified in API-610 are specified for ISO-2858 pumps in ISO-5199, which is adopted in IS-13537 . But ISO-2858 (IS-13518) can be read with or without ISO-5199 (IS-13537). It is for the purchaser to tune up the purchase specification to be appropriate and economical for the application. In America also they have ANSI-B-73.1, which is more akin to ISO-2858 and serves general purpose requirements which do not need specifications as stringent as API-610.

Q.    We are a company dealing in various torque measuring instruments. We     would like to know the importance of torque measurement in Various     Pumps.

A torque measuring instrument is important only where the measurement of torque is needed. Pump-users do not need measurement of torque. They would need measurement of parameters of pump's output, i.e. primarily flow and differential pressure. So a torque measuring instrument is irrelevant to pump-users. For pump manufacturers measurement of torque is somewhat relevant only during pump testing, if one wants to segregate input to pump from input to motor. Input to motor is anyway electrical and is measured by electrical measuring instruments. If pump is connected to the motor directly, without any transmission unit or gear box in between, input to pump is output of motor. And this can be derived from known performance characteristics of motors. This implies indirect measurement of torque. A direct measurement of torque using a torque measuring instrument has the advantage of it being direct. But the instrument out not to intervene or interfere in the transmission of the drive. If it interferes, it will falsify the measurement. But torque or torsion is a phenomenon, which does not seem amenable to measurement without intervention. There has been some talk of non-contact type torsion dynamometers, which again, if all feasible, would be a relative measure based on a calibration technique. To such extent it becomes again an indirect measurement only!

In assembly shops, for critical mechanical assemblies, people do use torque wrenches. But torque wrenches are 'tools' and not measuring instruments. There, torque is to be set to a 'control' and not exactly to be 'measured'. Setting the wrench to a level of control would imply as much measurement. But that calibration is provided on the wrench. If it is set to a torque of, say, 40Nm, it would not allow the exerted torque to exceed 40Nm. But it would not tell whether the torque actually applied was only 20Nm or 10Nm or 32Nm. That would be measurement but that again is not required in the use of torque wrenches.

Design Problem

Q.     Please advise the criticality of machining process for between bearing     split case pump.

Since there are two halves of casing and both halves are castings which have cored portions for wearing rings, impeller, volute cross sections, stuffing boxes and bearings, the cored portions have to match over each other and all machining is to be done by putting the two halves over each other with the gasket also in place. All this machining is thus blind and yet needs to be done within tolerance, especially at the wear ring and bearing portions, which have H7 or H8 tolerance. This machining is best done on a horizontal boring machine with a support column to support the special boring bar with number of boring tools for the different bore dimensions from one bearing to the other. If the pump is to be provided with mechanical seals, the stuffing box portion also becomes a close tolerance machining. Some designers separated the stuffing boxes in "through bore" design. But that increases the number of parts and number of surfaces to be machined and access to the rotating element for repair and maintenance does not remain as easy. Different approaches, different ideas, different pros and cons.

Q.    I want some clarification regarding a problem in Tubewell Submersible     pump.

    We have a water cooled single phase Tubewell submersible pump with     motor core length above 400 mm. To support both ends of the shaft, 27mm     journal bearings are used. Distance between the bearing supports is 540     to 550mm. The inner diameter of stator is 50 mm and outer diameter of     rotor is 49 mm We ore using 24 slot stator stamping and 16 slot rotor     stamping. The total weight of (rotor stamping + rotor conductors + rotor     end rings) is around 4.25 kg. The diameter of motor shaft is 27mm.

The problem which we are facing in the motor with above mentioned configuration is described below:

We observe a vibration and noise when the motor is running at no load itself. No load current consumption is around 7A and it increases up to 18 to 20A if we are providing a pump load. It is observed that there is a considerable speed reduction (from 2850 RPM to 2100 RPM) during pump testing. When we dismantle the motor after running of few minutes, we can seen that about 40 to 50mm in the middle of the rotor stamping is severely rubbing inside the stator. Only around 60 degree out of total 360 degree on the rotor stamping periphery is rubbing.

We are now trying to rectify the problem thinking that this may be due to a mechanical unbalance. We got the magnitude of total unbalance of the rotor shaft as 21g. Kindly give your suggestions and clarification regarding my doubts related to the above issue.

Q.    Why the rotor is rubbing inside the stator? What may be the real reason     behind it?

Q.    Whether there is any electrical force or magnetic pull affecting the dynamic     balance of rotor?

Q.    Whether I am now proceeding in the right way to solve the problem?

Q.    In which way I can proceed if the problem persists after using a     dynamically balanced rotor?

Q.    Is there any requirement of motor shaft hardening; if am using a     dynamically balanced rotor shaft assembly?

Q.    Kindly give some suggestions regarding different ways to solve the     problem.

For a visual, draw two circles one of dia 50 and another of dia 49,the center of circle of dia 49 should be 0.5 mm away from the centre of dia 50. The diagram will tell you why the rubbing marks are only on a small portion of the stator ID. The problem is either due to eccentricity during machining of the rotor, which also be a cause for imbalance in the rotor or due to too much clearance at the bushes or some foreign particles or burrs on the stator ID. The rubbing marks on the stator and rotor would be prominent enough to distract one's mind from looking also at the bushes. Since the rubbing marks are at the middle of the core length, that too on a vertical motor, bushes could be okay. But it would be good practice not to miss that. Any foreign particles would have got rubbed off, when you open the motor after running.

Rubbing is rubbing, no-load or full load. In fact, high no-load current should be considered as the right and first symptom of the problem and should be investigated right away, without carrying on with testing on load. Carrying on further is only worsening the damage. Why do that?

Q.    We would like to provide a water filling nut on our 0.50 hp pump model.     Please advise whether it is possible. Our pattern maker views are that it     can't be provided. Please suggest us on this subject.

This is a problem of individual design and would need to be considered separately and in the forum.

Possibility of providing the nut needs to be checked by examining the design/drawing. A good designer should know what is possible in a pattern and what is not. It cannot be the other way round, that a pattern maker tells what is possible or not.

Q.    What should be the sizing of the priming pot for Centrifugal pumps with Negative     suction. Is there any L/D ratio. Can a pump intended for positive suction be fitted     for negative suction with a priming pot

Volume of priming pot should be about six to seven times the volume of air-filled suction pipe. By this proportion water in the priming pot will be able to entrain some 15% air by volume. Smaller volume of priming pot will need longer priming time and more seriously may cause dry running of pump. Pumps intended for positive suction mean pumps having high NPSHr. Priming pot provides a virtual positive head upstream of the pump's suction. And the priming pot will have to be tall enough to provide the needed positive head. This would often be impracticable. It would be more logical then to install the pump with pump submerged, as a sump pump or a vertical turbine pump

For any centrifugal pump what is the minimum required NPSH for flow rate 5.0 m3/hr and 3.0 Kg/cm2 pressure head. In my case I have 0.3 meter available NPSH for above flow rate of fluid at its boiling point, please suggest me. Is it possible to have double suction centrifugal pump.

Available NPSH of 0.3 m is too small and does not provide for even the minimum recommended margin of 0.5 to 1 m between NPSHr and NPSHa. The required flow-rate of 5m3/h is very small. In a double suction impeller each side would get only half of the flow-rate, only 2.5 m3/h. There would be no double suction pump designs for such small flow rates. I would recommend simply putting the pump down into a pit and thereby increase the NPSHa. you further mention the liquid being near to its boiling point. That is dicy. It would be better to cool down the liquid to ensure that it is not in the vapour state. otherwise the pump would get vapour-locked and would not do any pumping.

New Excise Policy

Q.    We are the subscribers of your magazine almost since beginning. We     would     categorically like to state here that your magazine is very good and     extremely useful for our business. Especially the topics you cover are     extremely good and     useful.

Sir, for your information, we are the contractors dealing in Pumping machinery in Maharashtra and nearby States. We execute water supply schemes on turn key basis. As we require some information urgently, we request you to kindly let us have your professional guidance immediately In this regard, we refer Circular from Govt. Of India wherein it has been stated that Excise Duty is exempted on all those material which has been used for Water Supply Schemes. We also refer recent Vote on Account and declaration placed by Hon. Finance Minister wherein he has given benefit of exemption of Excise Duty on all the Water supply Schemes. For availing this benefit, we have to just collect letter from Collector or Asst. collector or Executive Magistrate confirming that the material has been used for Water Supply scheme. On submission of the letter to manufacturer, the material billed for the schemes will be excise free.

For getting confirmation of the same, we approached central Excise Office, Nagpur. However, we were surprised to here from the authority that such exemption is not at all available for Pumping Machinery and such exemption is available only for machineries used for WTP including pipe-line. However, there is nothing written or specified as far as Raw and Pure Water machinery and allied material is concerned. We also approached some leading civil contractors from Nagpur wherein they had informed us that they are regularly availing this facility from other few states like Gujrat etc.

Q.    Can we avail this benefit or not? We are in a genuine difficulty of whether     to avail this facility or not. We therefore request you to kindly guide us in     the matter along with necessary details regarding Notifications/Circulars     from Govt. Of India and also from Central Excise Deptt. so that we can go     ahead accordingly. Kindly quote necessary Notification and Circular Nos.     for our records for further necessary action.

It may be of interest to refer to Central Excise notifications, Nos. 46/2002 and 47/2002. Vide these notifications exemption was given in Central Excise for indigenous manufacturers, since import duty was also made nil vide customs notifications Nos. 91,92 and 93 all of September 2002.

The matter was discussed during council meeting of IPMA in Nov 2002. If I remember it right, it was not to be a blanket exemption. But, as I understood and if I remember it right, the onus was with project authorities to get the project notified and the exemption was to manufacturers and/or importers only for notified projects. The notification covered "All items of machinery". This would obviously include pumps.

It may have to be also checked whether there is any change in status during vote on account. However, vote on account is usually for administrative convenience of carrying on Governmental expenses. So no major policy changes are supposed to be effected during vote on account.

Tech Quest

Turbo Driven Boiler feed Pump

We have installed electronic governor on Turbo driven boiler feed water pumps so as control speed with reference to varying boiler feed water load. What is the recommended operating speed range within which, we can operate the unit without vibration problem. Turbine/Pump normal operating speed is 12090/2983 r.p.m.

The speed range will be specific to a given rotating element, i.e. load-diagram, moments of inertia of the rotating masses, locations of bearings, etc. One needs to find the "critical speed" and avoid also its harmonics.

I want to know the steps taken in a pumping system to attain the desired operating point for the system when

(a) the pump is driven by a fixed speed motor and,

(b) when the pump is driven by either a variable speed motor or a turbine.

The operating point is the point of intersection between the H-Q curve of the pump with the H-Q curve of the system. Once a pump is set into a system, this will happen automatically. But if the operating point, which happens automatically is not the 'desired' operating point, one has to modify either the pump curve or the system curve.

There are two ways to modify the pump curve -

1) Change the speed of the pump

2) Change the diameter of the impeller of the pump

3) The system curve can be notified by modifying the system. This is usually done either by changing the setting of the delivery valve or one can change it also by revamping the system by changing the pipe-sizes and/or layout of the piping.

4) If the suction conditions in the system are prone to cause the pump to cavitate, modifying the system to eliminate cavitation will also modify the pump curve from a cavitating condition to non-cavitating condition.

5) For changing the speed of the pump (option 1 above), changing the driver from an electric motor to a turbine will often become changing from a low-speed driver to high-speed driver.

Such change is possible even by using a gearing or pulley mechanism between the pump and the motor. But at increased speed the pump demands higher power input. So, it becomes important to check whether the motor has adequate margin in power. No such caution is needed if "desired" operating point is obtainable by reducing the speed.

For determining the required speed at the "desired" operating point, say

(Q",H")one needs to find the point (Qo, Ho) on the pump curve H = a*Q^2 + b*Q + c which also is a point on the parabola through the origin and (Q", H"). The equation of this parabola will be H = k*Q^2, where k = H"/(Q")^2.

Since (Qo, Ho) is to be a point both on

H = k*Q^2 and H = a*Q^2 + b*Q + c

to find (Qo, Ho) one needs to solve the quadratic (a-k)*(Qo)^2 + b*Qo + c = 0

Actually all the mathematics starts with knowing the values of the co-efficients a, b, c for the pump curve H = a*Q^2 + b*Q + c This is not difficult, if one knows three points on the curve, say, (0, Hso), (Q1, H1) and (Q2, H2) and solves simultaneous equations. A simpler way to do this is to plot the pump curve in an Excel spreadsheet and fit a 'trendline', setting also the option for the display of the equation of the polymonial of degree 2.

Submersible pump
Vs Turbine Pump

In our Municipal Corporation for raw, water pumping for 64 mtr head and discharge 1250 m3/hr capacity, we are considering procurement multistage (2 or 3 stages) turbine type pumps.

For the some application, one of pump manufacturer has very strongly recommended single stage submersible pump with motor submerge in the water.

This being is critical application, we request the views of your technical expert Mr Abhyankar, in respect of the following parameter, with respective above two alternative.

1. Efficiency of single stage pump with respective multistage turbine pump.

2. Bearing radial and axial thrust and bearing life.

3. Track record for similar application for above duty point else where in India of single stage submersible pump.

4. Your technical advise.

5. Cost comparison.

Basically, you are looking for pumping 1250 m3/hr of raw water across a total head of 64 m. The question raised by them is whether they should opt for vertical Turbine Pumps, which are quite common in municipal water supply pumping or venture into a single-stage submersible pump.

One thing to be always borne in mind by a pump user should be to consider the overall efficiency of the pumpset, inclusive of the motor efficiency. A pump user pays for energy consumed, which is influenced by the overall efficiency. By this consideration alone a submersible pump would get ruled out, because a submersible motor can never have an efficiency compatible with that of a surface motor.

A submersible pump would also need greater depth of the sump, in turn more excavation, so higher initial cost of constructing the pumping system.

If space is not a major constraint, I would also like them to check with another option of using horizontal split casing pumps. If one can have a dry pit alongside of the sump, priming of the pumps would also not be a problem. Maintenance activity is definitely very comfortable with these pumps, far more comfortable than with VT pumps.

The suggestions about checking construction costs, energy costs and maintenance really summarize into doing a Life Cycle Cost Analysis and take all possible options into account instead of restricting the thought process to one or the other type of pump at the planning stage.

This question I found in one of the IIT question paper. I could not get to the roots of the answer. We have a centrifugal pump with a 400 lpm capacity & 32 mtr. Head. An urgent requirement arises which need 500 lpm capacity with 50 mtr. Head. Without changing impeller, what are

The options available to meet this requirement? One option is to Change motor, what are the other supportive options available?

This is an important question, please give early reply,

The ratio of discharges, 500 lpm / 400 lpm is 1.25.

The ratio of heads 50m / 32m is 1.5625, which is square of 1.25.

Obviously the question is based on affinity laws, which are similar for change in diameter or rpm.

So another option to change in diameter is change in rpm, to 1.25 times higher speed,

MOC for Abrasion & Corrosion

I would like to know the difference between abrasion and corrosion. What type of M.O.C for impeller and shaft is suitable for abrasion and corrosion?

Also let me know the selection parameters for Impeller and shaft M.O.C.

One commonplace example of understanding the difference is water laden with sand. Sand, per se, is not corrosive, but it is very abrasive. Conversely acid with no entrained solids, clear acid, will not be abrasive but highly corrosive. Sea will also be corrosive. But corrosion due to sea water is due to its alkalinity whereas corrosiveness of acids is acidic in nature. MOC for corrosion resistance has to take into consideration whether the corrosiveness is acidic or alkaline.

Abrasion is also of two types. Abrasion due to fly ash in power stations will be from fine particles moving too close to the surfaces and abrading the surfaces. Abrasion due to sand particles or coal particles will be due to the particles hitting hard on the surface and bouncing back and hitting repeatedly. This is rather erosion than abrasion. So nature of abrasive wear depends upon the angle of incidence of the particles w.r.t. the surface. Usually hard surfaces would take abrasive wear better and resilient surfaces such as elastomer-linings would take the erosive wear better. But this is too much of a thumb rule. One needs to study the wear patterns and refer to the data available in handbooks.

Utility of Mechanical Seal?

I would like to know the basic criteria for choosing Mechanical Seal in place of gland packing.

Basically it will be cost-benefit analysis. The benefits have to also take into account benefits of environmental considerations. Where the leakages are likely to be hazardous, even safety considerations become important. For example, for radioactive leakages, even mechanical seals would not give adequate sealing. One may have to adopt zero-leak constructions as available in canned motor pumps or magnetically coupled pumps.

A golden mean between commonplace gland packing and mechanical seal is also available in the style of "injectable sealants", claiming the leakage sealing being as good as the mechanical seals and installation being as simple as rope packing.

Some people also offer a hydro-dynamic sealing with an expeller impeller at the back of the centrifugal impeller. The expeller impeller expels the leakage back into the pump and does not allow leakage to emit into the atmosphere.

Mechanical seals and injectable sealants claim to be bringing down also the energy consumption. That also should be a point in the cost-benefit analysis.

So, shaft-sealing has really become a multiple choice option.

Velocity triangle for Impeller

Here I am with two basic questions,

    Q. Can I know what is velocity triangle for Impeller?

    Q. What is its significance in the performance of the pump?

Waiting for your valuable response.

The questions would be of interest to anybody interested in pumps.

There are two prominent velocity triangles with centrifugal impellers one at the inlet and the other at the outlet. The flow is supposed to be travelling radially from inlet to outside diameter. This direction is similar to the line of longitude on the spherical surface of the globe, from one pole to another. Hence the direction is called as the meridional direction and the velocity is called as the meridional velocity. Next, inherent to the whirling of the impeller there is the circumferential velocity at every point on the blade. There would be a relative velocity between these two velocities. So the triangle brings forth the three vectors in relation to each other. The cross-section of the passage between the two blades and the layout of the blade causes the triangle to affect the vectors at every point along the blade. So there are virtually infinite number of triangles. But for a net result one would focus primarily on the inlet and outlet triangles. Yet, the layout of the blade and the total length of the blade will also vary depending upon how the angle between the relative velocity and whirling velocity, usually called as the angle Beta, is varied from its value at inlet to its value at outlet. Longer the length of the blade there will be better "Guidance" of the flow between two blades. At the same time, longer the length, the vane passage friction will be higher. So the trick is in striking the golden mean between guidance and friction. Obviously the pump-performance will be substantially influenced by how the trick is handled. That is why, it is always said, pump design is quite an art than just theory.

Also sir requesting you for a case study or material which explains about the meridional direction and meridional velocity.

Meridional velocities are assumed in the design procedure as outlined by Pfleiderer and Stepanoff. The notation for meridional velocity is Cm' for inlet and Cm" at outlet. Mr. Stepanoff in fact put forth empirical equations as Cm' = k' *sqrt(2gH) and Cm" = k"*sqrt(2gH) and put forth curves for recommended values of k' and k" vis-a-vis design specific speed.

1. What is the difference between centrifugal & Axial pumps?

2. In what way MOC selection criteria for casing and Impeller will be depended?

3. How to check the clearance between casing and impeller?

1. What is the difference between centrifugal & Axial pumps?

Axial flow pumps are one type of centrifugal pumps. specific speed in metric units will be >75

2. In what way MOC selection criteria for casing and Impeller will be depended?

Major consideration will be corrosion-resistance. For impeller, there will be also the consideration of cavitation-resistance and maximum whirling speed. Another consideration for impeller is of manufacturing feasibiity. This consideration is also relevant for casing. But with impeller the intricacy is more critical, especially for castability.

3. How to check the clearance between casing and impeller?

I would feel a fit from a tolerance of d11 on hub of impeller and D11 I on ID of wearing ring would be appropriate. I am writing this from memory. Please check whether this would work well.

How to determine Minimum stable continuous flow & Minimum thermal continuous flow. Is there any standard which explains about these parameters & the determination in detail

Minimum stable continuous flow is to be read on such H-Q curve which is unstable. In unstable characteristics, Hmax is greater than Hso (Head at shut-off). In such case, Minimum stable continuous flow will be where a horizontal thru' Hso will intersect the H-Q curve of the pump.

Minimal thermal continuous flow is that flow, when the liquid will experience churning caused by internal re-circulation. This happens because, the cross-sections of the hydraulic passages prove to be too large for the amount of flow to be carried. The designer designs the passages ideally for the design flow. At flows less than the design flow, the passages are not ideal. This is also one reason for the drop in efficiency at flows different from design flow. The effect becomes accentuated at flows less than Minimum thermal continuous flow. Churning of the liquid causes temperature of the liquid to also rise. This in turn raises the vapour pressure of the liquid. In turn the available NPSH gets affected. By all these considerations the curve for NPSHr v/s Q becomes uncertain. So, manufacturers show NPSHr curve only ahead from Minimum thermal continuous flow.

Obviously both Minimum stable continuous flow and Minimum thermal continuous flow are to be recommended by the manufacturer and cannot be obtained from standards.

In API-610 one finds a mention of continuously rising characteristics, that means a stable characteristics, i.e. where Hmax is only at shut-off. To be more mathematically correct, for a stable characteristics, the point of maxima is not in the first quadrant.

We have centrifugal pump having capacity of 60 m3/hr and we face problem of priming. we want to put self priming chamber. the suction pipe line length is 4 meter and have 80 mm diameter. can you suggest chamber size or formula to count same?

Volume of priming chamber should be about 7 times the of the suction pipe. Higher volume is helpful for to happen as much faster.

Yield of Borewell

What does it mean that a bore having 2 inches or 3 inches of yield. Please send us the chart.

Yield of bore coloquially called as 2" flow or 3" flow implies rates of flow usually experienced with pipes of these sizes. Technically speaking the rate of flow would be area*velocity. For a given pipe size as 2" or 3 " one would get different values of flow-rates for different velocities. I guess the colloquial talk assumes an average velocity of 3m/s. I need to cross-check whether this is okay.

Re-using Oil

We have 800 Ltrs. Used gear oil. Can we reusing the same after Filtration or other way you suggest.?

Please suggest & send Information for Re-Using of Gear Oil.

Reprocessing of used gear oil is a technology by itself. One needs to check whether the contaminants have affected the chemistry of the oil and the corrosiveness of the changed chemistry, apart from the lubricity and viscosity of the oil. The reprocessing will have to redeem the oil to original chemistry viscosity, lubricity, etc. Filtration alone may not do that.

Tech Quest

I would like to ask questions Regarding the Vacuum as per the Follows.

01) What principle use for vacuum pump.

02) Why the water Ring Vacuum Pump give 710 mm Hg vacuum only.

03) Why Ejectors inst on the top instead of Ground Floor.

04) What is the best way to check the leakage in Vacuum System's

All vacuum pumps are mostly positive displacement pumps. The question is of focus. When the focus is on what the pump is doing on the suction side, you are looking at a vacuum pump. When the focus is on delivery side, it is the pump. Even vacuum pumps on the delivery side develop some pressure.

All positive displacement pumps, the pumping action is within a close-clearance envelope. In liquid ring pumps, the close-clearance envelope is provided by the liquid ring. Since it is very essential to provide and maintain this envelope, there has to be continuous replenishment of the liquid.

Since tightness of the close-clearance envelope is not so good in liquid-ring vacuum pumps, they would develop vacuum only of the order of 710 mmHg In ejectors, also and, rather called as eductors, there has to be a high pressure feed to the inlet of the nozzle. Pressure being pressure, it will act, wherever the eductor is installed. Location or level of installation is according to what is suitable for the plant. If the vacuum is being disturbed, the gauge should show.

I want to know the information about the failure analysis of thrust bearing used in vertical turbine pumps. Please clarify with which organization you are working. For failure analysis of any equipment or component, one has to analyze the pattern of failure - whether sudden i.e. instantaneous failure or prolonged failure.

One has to also analyze whether the failure is due to improper monitoring of operational and maintenance parameters or failure due to improper design and construction. The objective of failure analysis would be to derive guidelines to prevent recurrence of the failure. To be able to derive proper guidelines, one would also have to do ROOT CAUSE FAILURE ANALYSIS. Analysis of prolonged failures. in particular, would need log-book data of how the equipment has behaved under what operational and maintenance parameters. If such data is not available, the exercise of failure analysis would get bogged down at the starting phase itself for want of adequate data. And, unfortunately this is often the case!! You will appreciate that a question such as yours will not have a simple, short answer. Your question itself is too short.

If I want to run centrifugal pump (bpo) with engine driven (application: fire fighting pump) then what is the margin to consider to select HP of the engine (i.e. 273cu.mtr/hr at 70 mtr. head efficiency of the pump-75% rpm-1800, then what is the recommended HP of the engine?)

With 75% pump eficiency f or 273 m3/hr and 70 m head, shaft or Brake HP works out to 70. Bur a fire-fighting pump is required to give minimum 65% head i.e. 45.5 m at 150% discharge i.e. 409 m3/hr. assuming pump efficiency will be about 65% at this duty point, shaft HP works out to 78 Hp. By this 75 or 80 HP should be okay. But engine Hp is to be further decided, also considering the altitude at the place and ambient temperature conditions. One needs to consult the engine manufacturers on this.

Please inform me about materialogy of parts, what is difference between SS 3l6,ss410& SS 316L, and also give me detail information about EN8,CI,Bronze,2%NICI,CA15,CS,CF8M,LG2 material.

SS 316 and SS 316L are austenitic grades of stainless steel generally having 18% Cr, 8% Ni and 2% Mo. 316 L has low carbon, as low as 0.03%. Hence the letter L. Grade CF8M has identical chemistry with Carbon up to 0.08% and the designation is for this chemistry obtained in castings. Designations 316 and 316 L are for materials in wrought forms, i.e. bars, plates, sheets, wires, etc.

En8 is commonly used wrought carbon steel for pump shafts. CS stands for carbon steel and does not denote any specific grade

CA 15 is 13% Cr stainless steel in cast form. It has moderate corrosion-resistance. But, since it is martensitic in microstructure it can be hardened, which makes it good for wear-prone surfaces such as shaft-sleeves.

2% Ni CI is alloy cast iron with 2% Ni. Such content of Ni is seen to make the material moderately corrosion resistant, especially in sea water applications.

Bronze and LG2 are copper-based non-ferrous alloys. Copper-based alloys are generally in 3 groups - Brasses, Gunmetals and Bronzes. Brasses are copper-zinc alloys; Gunmetals are Copper-Zinc-Tin alloys. LG stands for Leaded Gunmetal. The name Gunmetal seems to have been derived from its use for making guns in very very old times. Bronzes are also Copper-Zinc-Tin alloys with further alloying with Phosphorus or Manganese or Aluminium, hence Ph. Bronze, Mn-Bronze or Al-Bronze. Copper alloys were quite popular in old times, for equipment on ocean-going vessels, So much so that some grades got the name as Naval Bronze

How good is India compared to Western Pump makers for making pumps for process industries like Hydrocarbon Processing?? Do we still import pumps from the outside for refineries? Or are there some good, capable Indian players who can supply these equipments? I have heard of only Kiloskar Bros and Bharat Pumps & Compressors Ltd.

How good (technical capability) are the Indian pump makers compared to the ones in the developed countries, for industries like refining?

In hydrocarbon processing, standards for pumps are critical duties are of API, API-610 for centrifugal pumps. There are good number of manufacturers of pumps as per API-610, apart from Kirloskar Brothers and BPCL, This forum is not supposed to become brand-promotive. So other makes/brands are not mentioned here.

Q1. What will happen to a pump if its suction end and delivery end are joined?

Q2. Is it worth to change the casing of the pump than to change it completely if it has been seriously affected by cavitation?


Q1. What will happen to a pump if its suction end and delivery end are joined?

The pump will work in a closed loop or circuit. Both the liquid and the pump will get heated up.

Q2. Is it worth to change the casing of the pump than to change it completely if it has been seriously affected by cavitation?

The right approach is to do Root Cause Failure Analysis (RCFA). If it is established that the damage is due to cavitation, the reasons for cavitation have to be investigated. Replacing the casing or impeller or even the pump casing may not eliminate cavitation. New casing or new impeller or new pump may again suffer cavitation.

In a cooling water system, 2-pole pumps were suffering from cavitation. However,4-pole pumps for the same duty did not have the problem. NPSHr of 2-pole pumps was very high, whereas that of 4-pole pumps was low.

In another system, changing plug valves by ball valves reduced friction loss on the suction side and cavitation got eliminated.

Compressor Pumps

Q. What is the working principle of compressor pumps?

There are two types of pumps which use compressed air. One is Air-Operated Diaphragm pumps (AODs). Here air is used to activate and operate the reciprocating mechanism, to make the diaphragm pumps work.

The other use of compressed air is in jet pumps. Here compressed air is fed to the inlet of the nozzle. Across the nozzle, the pressure energy of compressed air gets converted to kinetic energy. Alongside the pressure energy become vacuumous. This helps the liquid in the bore to be sucked to the vacuumous condition at the mouth of the nozzle. The high velocity of air also helps to carry the sucked liquid to be lifted to the suction of the main pump.

In conventional jet pumps, a tapping drawn from the high pressure delivery side of the pump serves the same purpose as of the compressed air.

Whether with compressed air or with tapping from delivery side, the jet pump is never an energy-efficient way of drawing liquid from a depth. Jet pumps with compressed air would not need tapping from the delivery side. So, seemingly more discharge may be available from compressed air jet pumps than from commonplace jet pumps. But the carrier air is bound to occupy some cross-section of the fluid flow. To such extent liquid flow will be again less.

In commonplace jet pumps, the tapping from delivery side denotes amount of liquid being recirculated and energy being consumed in such recirculation. Even in compressed air pumps, energy of compressed air ought to be added to find the efficiency of the pump, which again will be correspondingly less.

Q. How pump with damage if run dry, which parts will be affected & for how long domestic pump can run dry without being damaged?

2. If we fully close the valve on delivery side, how pump will be affected? whether it will damage?

Dry running of pumps has different connotations. By one connotation, pump running dry means no liquid coming into the pump, i.e. pump not getting wet at all.

You have yourself sort of graded the question, making it into two parts. Let us analyse the logic step by step.

If no piping is connected to the pump, neither suction piping, nor delivery piping, and the pump is run in open air, pump runs dry, then the pump also is not doing any pumping work. Then no damage would happen.

If only delivery pipe is connected, and no suction pipe is connected, the pump would yet be running dry. The pump will work as a blower. Pumps are capable of working as blowers. No damage should happen.

If both pipes are connected, liquid level in suction sump is below the pump and pump is run without priming. The air inside the pump will keep recirculating. Pump will not do any pumping. No damage would happen.

As above, if both pipes are connected, liquid level in suction sump is below the pump and the delivery valve is kept closed, pump is run without priming. The air inside the pump will keep recirculating. There will be no escape route for the circulating air getting hotter. This would damage the pump, first and foremost the sealing components, then the bearings, then the rubber spider of couplings, alongside, lubricating grease or oil and so on. The pumping components themselves may not suffer damage, unless, the running clearances, especially at the wearing rings are too close and there is differential in thermal co-efficients of the mating components. The pump may suffer seizure and the whole pump will get damaged, the shaft getting twisted and what not. If the temperature travels to the motor, the insulation will weaken and the windings will suffer short circuit and the motor will burn.

By another connotation, if the pump is as above i.e. if both pipes are connected, liquid level in suction sump is below the pump, this time pump is primed, run. But, sometime during running the delivery valve is closed. Again the liquid inside the pump will suffer recirculation, will get heated up. This heat build-up often tends to be serious enough to cause all the damage noted above. Because there is liquid inside the pump, the pump is not dry as such. But the running of the pump becomes akin to dry running. With the heat build up, the liquid may suffer evaporation and the running of the pump will be more akin to dry running.

There are a number of ifs and buts in all the above analysis. It becomes no use to discuss the 'if's and 'but's after the damage has happened. People who claim longer dry running capability in their pump designs, would have to have more liberal clearances, materials of higher creep characteristics, higher class of insulation, distinct and effective cooling arrangements, maybe radiating fins even on bearing housings of pumps apart from those on the motor body.

Dry running capacity by virtue of liberal clearances would be at the cost of pump efficiency. Dry running capacity by virtue of better materials and/or with fins will be at higher material cost! "To get something, you have to sacrifice something." You need not 'get' dry running capability, if you can ensure operating practices with your pumps, such that the pumps WILL NOT RUN DRY.

In case of progressive cavity, helical rotor pumps, which have an elastomeric stator assembled snug over the rotor screw, all damage due to dry running will be "instantaneous"!!! The elastomeric stator being assembled snug over the rotor is essential feature of the pump construction. So, these pumps simply cannot be imagined to be built with any dry running capability.

Pump Efficiency

Q. How to improve Pump Efficiency? In all pumps. Please explain briefly. I want to know how can I identify the H.P of Motor in the field.

There would be two points of view regarding improving efficiency of a pump, one, in the field and the other, at manufacturing.

Improving efficiency in the field is the same as using a pump efficiently, which in turn is same as energy conservation in pumping. A comprehensive note on this can be found at www.pumps.org There, seven ways for saving energy in pumping are detailed. Briefly they are as follows.

1. Design systems with lower capacity and total head requirements. Do not assume these requirements are fixed.

2. Avoid allowing for excessive margin of error in capacity and/or total head.

3. Despite the tendency to emphasize initial cost, you will save in the long run by selecting the most efficient pump type and size at the onset.

4. Use two or more smaller pumps instead of one larger pump so that excess pump capacity can be turned off.

5. Use variable-speed drives to avoid losses from throttle valves and bypass lines, except when the system is designed with high static heads.

6. Use pumps operating as turbines to recover pressure energy that would otherwise be wasted.

7. Maintain pumps and all system components in virtually new condition to avoid efficiency loss.

Efficiency in manufacturing is through better design, better surface quality, especially of hydraulic surfaces of casing and impeller, close running clearances, least leakage, less wear at bearings, proper dynamic balancing and alignment.

Q. Whether API-682 seals can be fitted into ISO- 2858 pumps? If so, which edition 1 or 2?

ISO-2858 is adopted in IS-13518. Standard on "Dimensions of Stuffing Box Cavities" is IS-11382. It suits pumps as per IS-13518, i.e. ISO-2858. ISO-2858 was also derived substantially from DIN-24256 and the cavity dimensions derived from DIN-24960. DIN-24960 can be considered as the 'father' standard for API-682. But Americans are still attached to FPS system of units, whereas in European and in ISO standards the approach is emphatically metric and wherever possible, even SI system of units. American also have started giving metric converted values in their standards. But basic or reference dimensions are in FPS units and metric values are only conversions. So, dimensions in IS-11382 derived from an ISO-standard (I do not have the reference ready at hand), itslef derived from DIN-24960 is metric in its spirit. So there would nor be exact matching. Typical example is of Nominal size 3", which becomes 80 mm nominal in metric, but the difference is large, 3" being 76.4 mm. Even 25 mm being metric nominal of 1" also gives a difference of 400 microns, which is a large difference for shaft diameter, since there has to be much fi Further to my reply I would like to add the following. For requirements of pumps not necessarily as per API-610, ISo-2858 seems appropriate. IS0-2858 is already adopted in IS-13518. Actually, conditions such as vibration limits, minimum nozzle loads and moments as specified in API-610 are specified for ISO-2858 pumps in ISO-5199, which is adopted in IS-13537. But ISO-2858 (IS-13518) can be read with or without ISO-5199 (IS-13537). It is for the purchaser to tune up the purchase specification to be appropriate and economical for the application. In America also they have ANSI-B-73.1, which is more akin to ISO-2858 and serves general purpose requirements which do not need specifications as stringent as API-610.

I am a regular reader of your magazine esp. the tech quest part. I have few doubts in pumps as below.

1) If pipe length for a pump has to be increased, is it better to increase in suction side or in discharge side and why?

It is always better to increase the length on the delivery side. Because if done on suction side it will be accompanied by another adverse effect of reduction in available NPSH, which is susceptible to cause cavitation. Actually increase in friction due increase in length can be avoided, even more than compensated by increase in size of pipe.

2) If the discharge side of the pump is connected to the suction side, then what happens to the pump performance? That is whether the discharge pressure keeps on added every time or the TDH will be reduced only to make the friction in the loop? What happens to the flow in the correct case?

By connecting discharge of pump to its suction, it becomes a closed loop system. Commonplace example of such a system is transformer cooling oil recirculation. With a closed loop recirculation, the temperature of the liquid will rise. So effective radiation of the heat is necessary. Cooling water circuit of automotive engines seems similar. But it is not exactly a closed loop, because the cap on the radiator is not really air tight. Any time if the suction becomes open to atmospheric pressure, it does not become a closed loop system. In such cases, connecting the discharge to suction becomes like the discharge pouring into suction sump. Pump always works in a system and its performance is in response to the system head.

3) It is said generally that "It is better to start the pump with discharge valve closed". But it is also said that "Pump should never run at shut off condition". So during startup is there any time limitation for the discharge valve to be in closed position (the case is for a manually operated discharge valve)?

Pumps should be started and stopped in a manner that the load on the motor or power drawn by the pump from the motor is the least. This helps to get the motor to run up to full speed. With pumps, the load on the motor is the least when Q=0. For Q=0, the discharge valve should be closed. As far as the duration of keeping discharge valve closed is concerned, the operation is similar to putting a manually operated star-delta starter from "star" starting into "delta" running.

4) Pumps cannot handle air as it is designed for a hisher dense fluids (eg. water). Sir, is it possible for compressors to handle water?

In special applications pumps do handle air or gas along with liquid. For example, sucker rod pumps in oil exploration. Sucker rod pumps are actually reciprocating plunger, positive displacement pumps. Most positive displacement pumps can handle air or gas along with liquid. This is possible with positive displacement pumps, because, in these pumps, whatever is displaced is a positive effect, no return of displaced fluid. That is why they are called as positive displacement pumps. In case of centrifugal pumps, the displacement does not become positive, because of the clearance between the rotating impeller and stationary casing. Actually impellers of centrifugal pumps also do throw the air centrifugally. But the displaced air keeps returning to the suction through the running clearances. Air would require only an excuse of a clearance and the running clearance is more than the needed excuse.

Actually centrifugal blowers are not much different from centrifugal pumps. And blowers do handle air and compress it, though to a relatively smaller compression ratio. The passages in impeller and casing are sized very liberally compared to those in centrifugal pumps, This helps to make the casing to behave as a receptacle of the displaced air and running clearances become a small fraction of the displaced volume, thereby reducing the influence of clearances on the positivity of displacement. In centrifugal compressors also handling and compressing air becomes possible because of close clearances.

Handling air is all a game of qontaining the influence of clearances on the positivity of the displacement.

Tech Quest

Myself C.S. Panisha, subscriber of Pumps India Magazine , I seek yr guidance regarding the peculiar properties observed in subzero temp(-50 deg)

We are facing peculiar problem in pumps i,e Up to -45 deg pumps is giving design flow and pressure beyond -45 deg pumps flow suddenly drops , herewith giving you the details of Pumps

Capacity - 30 Cum/hr

Head - 50 mtrs

RPM - 2850

impeller Dia - 204 mm

MOC -- CF8M

I Sincerely solicitate yr opinion in this regards

In order to respond, it is necessary to know what is the liquid handled by the pump.

It is unlikely that Pump has mechanical reasons for the abrupt drop in flow rates.

If the liquid properties undergo a change which affect the suction or discharge conditions - reduced flow will be the result.

The clue lies in the liquid property at minus 50 deg C vs at minus 45 deg C.

Crystallization, gelling - in case of thixotropic liquid or simply a sharp increasing in viscosity - all can contribute to blocked suction and/or discharge condition.

If the pump resumes normal pumping at minus 45 deg C it is almost certain that problem lies in the system/liquid.

At sub zero temperature metals turn brittle and one may experience a shear, especially in hardened parts. If the process is designed to handle minus 50 deg C, so would the pump !

May be the liquid itself is not according to the specs specified - check this out too!

Trust this helps to look for the reasons of low/no discharge from pump.

I would like to be clarified briefly on NPSH parameter, how to calculate easily as client can understand it ,and judge the offering pump is suitable for the location.

pump for negative suction

pump with positive suction, tank fully closed inside of hot water positive suction with tank in under vacuum.

What you are referring is NPSHa - available NPSH (Net positive Suction Head)

Pumps India has in the past covered this question. This does require an explanation with a speech background for every illustration for better understanding.

Simply stated it is the suction head available at the pump inlet - expressed in metres or feet (in USA)

Mathematically it is expressed as :

NPSHa = +/-Hzgeo+{(Pg+Pabs-Pv)/ng}+V^2/2g-Hf

Where Hzgeo - height difference from system datum level to the impeller datum (in case of centrifugal pumps) or

simply at

Suction flange level (if suction flange is vertical)

For suction lift use -Hz, when suction head use +Hz

Value in metres [m]

Pg = gauge pressure in Pa. [Pa = N/m^2 =10^5 bar]

Pabs = absolute pressure at datum level in [Pa],Pascals

Pv = Vapour pressure of the fluid at pumping temperature
[Pa]

n = Density of liquid [kg/m^3]

g = Gravitational constant =9.81[m/s^2]

V = flow velocity at suction [m/s]

Hf = Head loss in the suction piping (due to friction, sudden contraction etc) [m]

For vessels/tank open to atmosphere the value of Pabs is
1020 mbar = 102000 Pa = 10.2 mwc

For vessels under vacuum use the absolute values instead of 1020 m bar

Value of n = 1000 kg/m^3

Values of vapour pressures are available in reference books for most liquids at different temperatures, for water at 100 deg C it is

1.01 33 bar or 1013,3 mbar,

at 30 deg C it is 0.0424 bar = 42.4 mbar

at 10 deg C it is 0.01227 bar =12.27 mbar

For liquids that are volatile the vapour pressure values are significant even at room temperature.

In practice, to make the calculation simple pressures are converted to mwc or mlc (metres of water or liquid column),neglecting the denominator n*g= 9,810 rounded off to 10.000 thus allowing Pa to be divide,C by 10,000.

This simplifies a value like 1020 mbar = 102000 Pa to read as 10.2 mwc

Some other values you use commonly are vacuum expressed in mm of Hg (mercury). There are nuances between the terms bar, atmosphere(technical) [at] and atmosphere (physical) [atm].

Values expressed in mm of Hg will be = 760 against [atm]

= 750 against bar

= 736 against [at]

Commonly 760 mm of Hg is used as a standard value. The error is negligible if you use this in place of 750 or 736.

May be in one of the forthcoming issues of the magazine, we will try to cover this elaborately for all readers.

The confusion/misunderstanding exists even among the consultants and Pump engineers.

Today you have several websites offering detailed explanation - some with animation as well. Do browse the net !

You can also read about NPSH in any "Pump handbooks" and in "Centrifugal Pumps" by Igor Karassik.

My query is about how can we decide actual power (H.P.) of a motor?

Assuming you have no nameplate to guide you, you should be at least sure of the voltage for which it is designed. In running condition, you will need to connect the motor to some kind of load. Measure the speed -with a tachometer. and measure the torque with the help of a torque transducer. Speed (rpm) x torque (kg.m)/955 = kW, if you use Nm units for torque the factor will be 9550 instead of 955. [for Hp. use factor 712 or 7120 depending on the units used for torque) Most of the users will not have torque transducer and variable load like pump. It is easier to measure torque if you have double ended shaft motor which can be mounted of external bearing supports and torque measured with a torque arm of known length and weighing scale that can measure force applied at the end of the arm (wt x arm length -[measured from shaft centre] will give Torque in Kg.m If you have only the rpm measured and have no other data, motor shaft dia will be an indication to
HP. For this refer to dimension catalogue of any standard motor manufacturer. Still simpler, measure the height from base of the motor to the centre of shaft (we call this as centre height). This measurement in mm will correspond to 'Frame Size'. Refer any manufacturer's table corresponding to the speed under this frame size to read HP. I am assuming the motor is foot mounted and is manufactured to IEC standards. For your own make motor or odd motors like submersible motors. torque and speed

measurement is the only sure way of HP or kW. The relation between these two: HP x 0.746= kW Yet another method is to measure current and voltage applied. Product of the two x 1.73 gives apparent power in watts. Dividing this by 1000 will give kW. Further multiplying this by power factor and efficiency will give power in kW.

For example a motor on 415 V 3ph consumes 27.5 per phase, power factor at 100% load is 0.85 and efficiency at 100% load is 89.4% The kW rating will be =415 V x 27.5

A x 1.73 x 0.894 x 0.85 = 15.0 kW

For single phase motors you should drop 1.73 factor used above (square root of 3)

What are types of pump, what is volute casing pump, what is no volute casing pump, types of pump according to motor using, pump components

Question requires very large space to answer in details.

Briefly, volute casing in a centrifugal pump is a static part that converts the velocity head of liquid pumped into pressure or 'head'- kinetic energy imparted by the impeller to the liquid is converted in to potential energy.

Volute refers to a geometric curve that is generated by the end point of a straight line when the line moves tangentially along the circumference of a circle without skidding.

A vast majority of the centrifugal pumps have volute casing - easily identifiable by its external appearance. Variations include a double volute casing & split casing pumps.

The function of volute casing is also carried out by 'diffuser' which uses guiding vanes and is commonly used in multistage centrifugal pumps.

Concentric casings also are used for certain applications despite the fact that the radial forces acting on the impeller are of higher value.

Side channel, regenerative (peripheral-turbine), recessed impeller pumps - also called vortex pumps, are some other types which are often clubbed with centrifugal pumps although their action of pumping is different from conventional centrifugal pumps. For want of space illustrations are avoided here but you will find these in most articles in the magazine as well in the advertisements of manufacturers Motors don't decide the pump design except for the Monoblock pumps where the motor shaft itself supports the impeller.

Since a vast majority of pumps are driven by electric motors - the motor speed becomes an important factor in selection of Pump size. Direction of rotation is important as well but with most applications allowing usage of bidirectional motors it is a less significant factor.Spares, however, must be ordered with direction of rotation in mind (Especially the impeller and shaft).

For the other part of the question- I would refer you basic books like Centrifugal Pumps by Igor Karassik, Pump Selection and Application, Pump operation and Maintenance by Tyler Hicks that are very illustrating.

Basic principles of centrifugal type of pumps

Basic principle of Centrifugal Pumps - as the name suggests it has do with centrifugal force acting on the liquid.

If a liquid enters at the centre of a rotating disc it will be subjected to a force acting from centre to the periphery of the disc in a radial fashion. While passing from centre of disc to periphery it acquires velocity. This kinetic energy of fluid is converted into head with the help of a volute casing or diffuser.We call the disc as impeller which has vanes to pump liquids efficiently.

You have many websites that have animated explanations for such principles! Do visit them!

Is there any limiting value of Impeller Diameter to, a single stage overhung pump running at 2900 rpm. What are the main reasons that limit the Impeller Diameter apart from disk friction losses? Please also let me know if there is any(certificate) course available on the trouble shooting and Industrial application of centrifugal pumps, compressors and other rotating equipment. Thanks in advance for your cooperation.

The centrifugal forces acting at the tip of the vane/impeller periphery and the physical property of material used for construction of impeller limit the tip speed to 40 m/s. This would translate as 264 mm as the limiting dia for 2900 rpm

In most makes of centrifugal pumps you will observe this, Exceptions are manufacturers like Sundstrand who offer special metallurgy and design for their impeller that can operate at speeds in excess of 15000 rpm. But even here the limits are within 50 m/s.

Singapore Tech University should have some courses. In USA there are many professional institutions that offer such courses with hand on training. You will get them on the net for asking !In India there are certificate courses but there is very little practical training imparted.

Why does API 610 recommend specific margin i.e. 25%, 15% & 10% for selecting motor rating over pump rated bkw?

The basic reason for specifying motor rating above that of the BkW of Pump, API or no API, is that the duty conditions themselves are not rigidly fixed. e.g. variations in discharge head, suction conditions, viscosity, temperature -all can affect the duty conditions.

In centrifugal pumps a lower head will result in shifting the duty point to the right of BE,P or original duty point. This will invariably reflect as higher BkW. A change in the S.G. of liquid directly affects the BkW. Basically to take care of these variations a margin in the Motor rating is desired. These values have been arrived in consultation with the users and manufacturers.

Tech Quest

Q. How pump selection is done for various application, what are the basic things to be considered, what are the basic things to be know before selecting pump.

For a good selection of a pump, knowledge of various types of pumps available is essential. Efficiency, reliability and ease of maintenance are important considerations for selecting a suitable pump.

-Liquid to pumped and its properties such as sp.gravity, viscosity, its effect on metal/no metals used in the manufacture of pump

- Flow rate and the total pressure to be developed

- suction conditions, requirement of self priming

- Application -requirement of constant flow, variable flow capability

- Special process requirements - non-emulsifying handling, solids handling capability

- Ability of pump to work at motor / engine speeds

- pumping temperature,

- Sealing arrangements (permissible leakage) are some important factors needed to be considered. History of satisfactory performance plays an important role. Based on the years of satisfactory performance, most of the consultants specify the type of pump needed. Broadly speaking there are two basic pump types : Kinetic pumps and Positive Displacement pumps.

Kinetic Pumps include all pumps that impart energy to the pumped liquid by virtue of speed (rotating impeller in centrifugal pumps)

Positive Displacement Pumps -which deliver constant flow regardless of pressure per rotation or stroke, where flow is directly proportional to speed (rpm or spm) Centrifugal pumps handle clean liquids as well as slurries. Large capacity and medium pressures are best suited to most, Special pumps like multistage type or high speed pumps can deliver high pressure as well but restricted to clean and low viscosity fluids.

Positive Displacement Pumps on the other hand are suitable for viscous fluids and high pressures if required. Reciprocating/Rotary pumps like gear, screw and vane are some well known types.

We reproduce a capacity and pressure related selection basis. By no means this is rigid. User/manufacturer's experience is invaluable.

Generally for oils and viscous liquids - positive displacement pumps are preferred. If viscosity is above 200 cSt - the choice is invariably a PD pump.

Where flow variation is needed, As in a process plant, large capacity and medium pressures are needed Centrifugal Pumps are the first choice.

Centripetal pumps, multistage pumps, submersible pumps, vortex pumps, peripheral pumps are some other kinetic pumps. In PD pumps you have Piston, Plunger and diaphragm are examples of reciprocating pumps.

Rotary PD Pumps are : Gear(internal and external type), Vane, screw, lobe, shuttle block and many other types. Most manufacturers use a software to select a given pump with the result, the expertise in manually selecting appropriate pump for a given job is getting increasingly lesser!

Your generation is lucky to have an access to various web sites to get information at the click of a mouse!

For selected reading on pumps, we would recommend at least the following two books:

1) Pump handbook - by Karassik and paul

2) Pump Selection and Application &

3) pump Operation and maintenance - two useful books by Tyler Hicks

Q. What is the difference in application and working of submersible pump, vertical turbine pump, horizontal centrifuge pump, and metering pump?

The pumps referred by you differ in construction and application.

Submersible pumps : There are two types - for borewell application or dewatering of mines with high capacity and head.

These have a pump motor combination : Pump on top and motor at the bottom. The intake is in the middle.sizes up to 6" are normal for domestic use, beyond this it is usually for large water supply or mine dewatering jobs.

Construction is similar to that of a vertical turbine pump : impellers, bowl assemblies and diffusers as the pumping elements.

Normal speeds are 2900 rpm. Multistage construction is required to meet high discharge heads.

Vertical turbines are for larger capacity. Smallest bowl assembly would be of 4" and meant for open well application.

Bowl assembly is at the bottom. Column assembly supports the bowl assembly and serves as the water passage. Drive shaft is centred and lubricated at intermediate flanged joints with cutless rubber bearings when handling clean water. In case of dirty water an additional oil tube serves to lubricate the shaft joints and isolate it from dirty water.

Motor is usually a hollow shaft type - taking care of pump thrust and providing the torque and rotation for the impeller.

Large capacity and heads from medium to high is the main application area. Water supply schemes, lift irrigation, storm water pumping are the main fields of application Industrial application includes duty liite ATF handling at airports..Larger sizes - known as mixed flow types provide larger capacity at moderate head- used extensively in thermal power plants for circulating and cooling water supply. Large lift irrigation schemes is another area of application.

The above described pumps must have at least the first stage immersed in water. They are not self priming.

Other submersible pumps have motor on top and pump at the bottom. Usually a monoblock construction - i.e. Pump impeller is mounted directly on the motor shaft, These have a volute casing like a normal centrifugal pump. Low to medium head with medium capacity is the strong point for their application : sewage ,storm water, tank draining and emergency dewatering of pits are some regular applications. Here motor is submerged and depends on water for cooling its external surface.

One more variety of submersible pump is a combination of volute casing & impeller submerged but the motor is located on surface. A vertical column assembly with rotating shaft - like in a vertical turbine pump are the other components of the pump.

Used for leakless pumping of solvents, industrial waste, municipal wastwe, molten sulphur and other chemicals.

Horizontal centrifugal pumps - the most common types- have an impeller and volute casing as the main pumping elements.

Driven by a shaft supported in a bearing housing. located at the surface. Pump can be with cantilever design or with impeller supported between the bearings as in case of a split case pump. Variety of liquids and a wide range of flow/head range is available. Process industry is the largest user of such pumps.

All the pumps described above are kinetic pumps - where the energy is passed on to liquid by impeller. Speed and impeller dia are important factors.

Metering pumps are - as the name states for metering or dosing of precise quantity of fluid regardless of minor changes in the discharge pressure. Usually, reciprocating plunger or diaphragm type with fluid ends suitably made from metal/alloys/polymers to withstand the corrosive action of liquid. Small capacity and high pressure capability are common. Dosing of chemicals for water treatment or for process are the main applications. Capacity can be varied mechanically - by adjusting the stroke length and in some cases by changing the strokes per minute (spm) -in some designs both can be varied.

Refer any pump handbook for details. Chemical Engineering Handbook by Perry has a good summary.

Hope this helps.

Rectangular weir is simple to construct : Width of channel Lw should be= 4 x H + Lc

Where H- the height of weir (m)
           Lc - width of weir (m)
           C - height of crest - the distance from the bottom of channel to the bottom of weir (m)
Channel width should not be less than Lc + 4 H
           C should be greater than 3xH

For flow rates up to 1 m3/s you could use a weir with width of about 1 m, height of weir 0.3m, total length of weir about 6 m

The formula for flow rate is

Q = 2/3 Cd x Lc x (2g)^0.5 x h^3/2

Where Q is in m3/sec

Cd is coefficient to be found by experiment

Lc -in m

h - height of water over weir measures at least 4H distance upstream of weir

g - Gravitational constant 9.81 m/sec2

You will get this data in any hydraulic book or if you have an access to the net, log on to "rectangular weir" on google search and you will have tons of material well illustrated.

Q. I want to know comparison between submersible motor and dry motor.

Dry and wet when referred to electric motors simply mean the design suitable for use in air (above water - not immersed) and immersed or submerged in the liquid-usually water.

Pumping element - impeller and casing may be in the liquid but the drive motor is above the liquid.This would be a dry motor.

When we refer to submerged pumps such as borehole pumps or drainage pumps, the motor is also under water. The design should prevent ingress of outside water into the pump. The cable entry is hermetically sealed and there is no cooling fan on such motors. These are cooled by external water/liquid which must act like a heat sink.

The term 'wet' does not refer to the oilfilled or waterfilled motors although they will invariably be used for under water (wet) applications.

Attached are illustrations of wet and dry motors (with pumps) for your info. Hope this clears your doubts.

Dear Mr. Rao,

The equation for mass flow based on density and velocity :

M(dot) = rho*V*A

Where m(dot) - mass flow kg/s

rho = density-kg/m3(kg/cubic metres)

V = velocity-m/s

A = area of cross section - m2 (square metres)

For gases rho(a) = Pa*M/Ta*Ru*Z

Where * denotes multiplying

Pa - absolute pressure N/m2

M - molecular wt Kg/mol

Th - flowing temp absolute (k) = Kevin

Ru - universal gas constant N-m/mol-k =8.3144..

Z - gas compressibility factor, usually 1 for STP. For other conditions, varies between 0.4 to 2 depending on pressure and temperature.

In the days of internet, you have several sites offering free calculator all that you do is fill in the data in the windows specified and you have the readymade answer.

However, we must always know the fundamentals to understand and derive the formula if negded.

Visit to www.oseco.com may be of interest for you.

Q. We have a soap industry, where in the process we have to transfer hot flowing soap to a height of 32'. We are using Rotodel gear pump of 1440 Rpm reducing it to 900 Rpm by a Pulley with a 10Hp motor but it is not very sucessful as there is a lot of wear & tear of pump. Kindly suggest a better pump.

Molten soap handling is a common application. The problem lies in getting the pump filled completely before the pumping action takes place.

Gear Pumps have been used in industry. The secret lies in reducing the speed to an extent that will match the pump output to that of the incoming flow. The input is assisted by gravity and by the differential (partial vacuum) pressure created by the pump.

You may have to try this out by using a variable speed drive to check at what speed the pump offers optimum performance. The process a bit complex.

At the start of operation the reactor vessel if full - having maximum static head for pump. As the time progresses.the static head keeps reducing thus reducing the flow rate to pump.

Usually the plant supplier is aware of this and provides the necessary changes in the drive.

In any case 900 rpm seems too high.Start with 200 rpm and step up slowly to check how the same pump behaves better. Shuttle Block Pumps (Rotary Piston)/Tushaco/ have been used by Godrej/Alfa Laval for similar application. Speeds are always well below 750 rpm.

Precaution : Do not use the pump during flushing with steam and water. A few minutes of use in water is equivalent to years of working in soap!

If affordable- the best alternative is to- use an external lobe pump (driven by a pair of timing gears). Tuthill of USA is well known for this. These pumps can be safely used during flushing with water.

In India,Fristam,Alfa Lava/LKM have designs similar to this but only in SS version.

Tech Quest

Residual Life of Pumps
Q.     Our plant is one of the largest LPG unit in the country. most of the pumps are 20yrs old. The fluid handled by different pumps varies from c2-c3, lpg, crude oil, utility water, effluent water etc; can you suggest evaluation methods for fitness for service, analytical & practical method for residual life of pumps for replacement; present operating parameters(mech. & elect) and condition seems o.k based on condition monitoring observation. can you suggest some web sites or books for residual life analysis of process plant dynamic equipment?
Any pump has distinct zones, the wetted zone, the shaft-sealing zone, the bearing arrangement, the coupling and drive. In the wetted zone the pump casing is a pressure vessel and the residual life will be related to the wall-thickness. Pumps handling corrosive and abrasive fluids will have fast thinning of the all-thickness. The important component in the wetted zone is the impeller and the clearances at wearing rings. Wearing rings are of course replaceable parts and clearances can be redeemed to "like new" condition. Cavitation, erosion and corrosion are three major items that will be eating into the life of the impeller. Shaft-sealing zone is always attended to, whenever the leakage exceeds permissible and residual life gets revived with every refurbishment. Bearing life for journal bearings is theoretically infinite, because journal bearings a film of lubricant is supposed to be maintained all through the working of the bearing. If that is provided, there should be no wear and hence infinite life. For anti-friction bearings, the pump-designer selects the bearings, considering the axial and radial loads and desired bearing life, applying also the recommended factors for severity of service. Condition-based maintenance should be giving good indication of the condition of bearings. Most of the components of the "rotating assembly" are replacement spares. All replacements would redeem the residual life to "like new" condition. If so, scope for residual life analysis brings focus primarily to the pump casing, which should be looked at as a pressure vessel, where wall thickness becomes the guiding parameter.

Dosing Pumps & Gland Packing
Q.     What are different types of Gland Packing? How does one determine their use?  Under what conditions does one use Teflon and Asbestos Gland Packing? We have four Multiflo(V K Pumps) PR-10 Model Pumps that are used for dosing Chemicals in HRSG. Two Pumps serve as LP Dosing pump discharging Hydrazine+water solution at a Pressure of 4kg/cm2. The other two pumps are used as HP Dosing pump discharging Trisodium Phosphate solution at a pressure of 42 kg/cm2. Can we interchange these pumps? Can a LP Dosing Pump now discharging 4 kg/cm2 discharge at 42 kg/cm2 since both the pumps are of same models?
Usually the design of the 'liquid end' of reciprocating piston pumps is related to flow-rate, (volume per stroke multiplied by number of rokes per minute). Volume per stroke is cross-sectional area of piston multiplied by length of stroke. Mostly in a given model of pump, these parameters would be same. If so, one should be able to interchange pumps from one duty to another, provided the flow-rate is same. Pressure would influence the wall-thickness, especially of the cylinder. If that also is same, one can interchange more confidently.

Various types of Pumps & Applications
Q.     Can you Please provide me the details of the following :-
1)     Definition & Details of Operations of Centrifugal Pump - Also Industrial Applications
2)     Details of Operations of Rotary & reciprocating pumps - Also Industrial Applications
3)     Operations Details of Parallel & Series operations - Applications and Purpose.
4)     Types of Vertical pumps and its operations
5)     What the Difference between a Centrifugal Pump & Reciprocating Pump
A complete and comprehensive answer to the question amounts to writing a wholesome book on pumps, because the question asks for definitions and details for centrifugal, rotary and reciprocating pumps, vertical pumps and for applications including industrial. I think some commonplace examples will explain the three basic classes of pumps. The pump used to inflate the tyre-tube of a bicycle is a reciprocating pump. A to-and-fro motion is a reciprocating motion and all pumps where reciprocating motion is employed to do pumping action are reciprocating pumps. Swelling and shrinking motion of our heart t inhaling and exhaling is a motion similar to a reciprocating motion and we have diaphragm pumps working similar to our heart. Diaphragm pumps are also a type of reciprocating pump. The mini monobloc pumps used in many households are centrifugal pumps. Centrifugal pumps are very very commonplace and can be found everywhere, on the dug wells of farmers, in multi-storey residential premises for raising water from ground level reservoir to the overhead tank, etc. Pumps at petrol pumps for filling petrol or diesel in vehicles are most commonly rotary pumps, technically called as rotary positive displacement pumps.

Q.    What are canned pumps?
The word "Canned" gets associated with pumps in two contexts - one, in respect of "Canned Motor Pumps" and another in respect of pumps handling volatile liquids like LPG, where a vertical pump has a can around it, deep enough to provide the NPSHr of the pump by virtue of the suction impeller getting as much positive head.

Q.    Wanted some information on Osmotic / Osmatic Pumps.
From the name of your company you seem to be involved with water treatment. Reverse Osmosis is a process for purification of water by forcing the raw water through osmotic membrane. Osmosis is a natural phenomenon, by which thinner liquid will travel towards a thicker liquid, even if the two liquids are separated by a permeable membrane. But, for purification, raw water, which is thicker, has to travel across the membrane. This is contrary to natural phenomenon of osmosis. Hence the raw water is required to be forced. And the process is called Reverse Osmosis (RO). The pressure required is quite high, depending upon the content of impurities. RO is supposed to be the process capable of giving purest water, not allowing even micro-organisms to pass through. Because of high pressures, RO was not considered feasible for domestic usage. So, domestic water purification systems were based at best on purification with Ultraviolet rays. However over the last 6 to 8 years domestic RO systems are also available. RO systems have been used on ocean-going vessels since many years. Removal of salinity from sea water to make it potable is again most reliably achieved with RO. Pumps for RO system are in the entire range of sizes from those for domestic systems to even municipal water supply systems.

Axial Float
Q.    How to calculate axial float in boilerfeed water multistage pumps.(with and without    balancing disc) please enlighten me with diagrams.
Axial float is caused by the axial thrust. There is axial thrust even in single stage pumps. It gets multiplied and hence becomes more pronounced in multistage pumps. Basically axial thrust is caused by a differential pressure across an area. In a closed impeller area of the backshroud between the hub dia and eye diameter experiences suction pressure on one side and impeller's discharge pressure on the other side. This would cause the rotating assembly to be thrust forward towards the suction. In case of semi-open impellers, there is a pressure gradient all the way from the outside diameter of the impeller to the hub diameter. In case of vertical pumps, the total weight of the rotating assembly would also add to the axial thrust. Calculation of axial float will then be related to by what amount the axial thrust would displace the rotating assembly. In most designs the displacement will be arrested by the bearing clearance. Even commonplace deep groove ball bearings are also available in reduced or controlled clearance version and are used in pumps with semi-open impellers. Axial thrust apart from being a regular aspect of a running pump, it becomes a transient phenomenon during starting and stopping of pumps and would cause an impact onto the feature arresting the displacement, viz. bearing clearance or seating face of balancing disc. Balancing disc employs the same basic concept of axial thrust, it being caused by differential pressure over an area, but in the reverse direction. An extempore thought comes to mind that designers can try reducing wear due to transient impacts during starting and stopping, especially during stopping by providing a tapping from beyond the non-return valve to the balancing disc. This may be possible even in existing pumps. Most pumps with balancing-disc are with journal bearings. Hence the balancing disc only serves as the displacement arresting feature. It is susceptible to wear. Hence all balancing disc have a raised portion ring near the outside diameter of the disc. The maximum wear allowance is the height of the raised portion. This is marked on the shaft for easy monitoring by the user for the wear on the balancing disc. Since balancing disc operates with the principle of differential pressure over an area, it is important to ensure the pressure differential. The parallel faces between the balancing disc act as an orifice and cause the differential pressure. So the moving face needs to be set to give minimum clearance, but not so minimum as to cause rubbing. Secondly the throttled pressure has to be tapped off. Otherwise the balancing disc area will just become a pressure vessel and the pressure will balance and there will be no differential pressure to balance the axial thrust. So the line tapping off the throttled pressure should not get clogged. Wear from the balancing disc is itself liable to cause clogging.

Selection of Pump for Agriculture Application
Q.    I noticed your detailed question and answers in the magazine from that I am willing to clear my doubts I am a farmer, farming about 60 acre with coconut and sapota groves. I am having part of the land near by a river and a part of the land too far. for that I planned to pass water by underground pipe line, the distance is 2 km approx. with a maximum head of 37 feet discharge height, suction head may be around 15 feet height. I planned to use the underground pipe water by dumping
the local wells before irrigation. for that, 1.What type of pump is needed.
2.     How much capacity is needed.
3.     Power consumption
4.     I planned to use PVC/HDPE in underground (which one is better), GI in open area.
5.     Give the pipe dimension throughout the distance Suppose: If I directly irrigate by using drip irrigation (the line provided for drip irrigation is 3" for the whole land)
1.     What type of pump is needed
2.     Capacity,- power?
3.     Pipe material of construction?
4.     Give the dimension of pipe throughout the distance
5.     Kindly mention the number of air vent valves to be provided.
You have mentioned that you find my answers 'detailed'. Your question has prompted really the 'most detailed' answer! Your question gave me the opportunity of bringing out also the cost, which the nation bears for providing power to agriculture. Thank you hence for your question. It is true that over the last 50 years our country has become self-sufficient in food. The nation is indebted to the farming community for that. But the nation has also borne very heavy cost for providing the power. And it has not been used any intelligently or efficiently by the farming community. There is rampant dishonesty, what with stealing of power! The power supply system cannot manage the quality of power against so much of stealing. And we grumble of very low voltages, fluctuations and trippings. Pump manufacturers end up designing pumps and motors to withstand these vagaries of power, hence that much costlier motors and as much inefficient motors. All these hidden costs finally usurp the pockets of the entire population of the country. Hope, you would find the appropriate methodology for making the right selection of pump. But you may have to do all the calculations all over, especially wherever the assumptions are not correct.
For a ready reference, the assumptions made are as follows.
1)     Water requirement for coconut is assumed as 100 mm every fifteen days.
2)     Water requirement for drip irrigation is assumed at one-fourth of the above.
3)     Frictional head for GI pipe only is taken for illustration. You may check for other pipes, especially HDPE.
4)     Other frictional heads for losses in foot valve, non-return valve, bends, etc. are not taken into account.
5)     The illustration considers only pumping river water to local wells. Another pump will be required for pumping from local well to the field.
6)     Assumption also is made to run the pump for eight hours per day.
7)     Percolation in local well(s) or yield of local well is assumed to be 5 liters per second.
8)     Total length for pipeline considered is 2 km.
All this procedure of selecting an agricultural pump and composing the system is also covered in Indian Standards IS-9694 Part 1 and IS-10804. Of these IS-10804 also gives ready data about frictional heads for different types of pipes.
For selecting a pump one has to first find out the required discharge and head. Discharge is a rate and not a total quantity. Suppose your 60 acres of land should be fed with 4 inches (100 mm) of water in one irrigation cycle, to find total quantity of water, which will have to be pumped, we would first get all data in proper units.
60 acres = 60*4840 sq. yards = 60*4840*9 sq. ft = 60*4840*9/3.28/3.28 sq. m. Then4 inches (100 mm) of water over this much area means 60*4840*9/3.28/3.28 *0.1 = 24293.6 cubic meters of water. Now you would not be able to irrigate all the 60 acres of land in one day. If you would irrigate say 4 acres per day, so that every acre would get water once in every 15 days, then 4 acres of land would need 24293.6/15 = 1620 cubic meters of water. This will vary. If every acre should get water once in every 10 days, 6 acres will have to be irrigated everyday. Then water to be pumped every day would be 2430 cubic meters. Again I just assumed water requirement as 4 inches in one irrigation cycle. It would be different for different crops and would be different for different methods of irrigation. In drip irrigation, the water requirement could be only one-fourth of that for furrow irrigation. You are planning for coconut and sapota. Now if sapota is to be a crop by inter-cultivation, it would not need separate feed of water. It would share from what is fed to coconut. You will have to check whether this assumption would be correct in drip irrigation, because, in drip irrigation, one feeds water to the root of one plant. So, if coconut takes all the water that is fed at its root, it would not leave any water for sapota. However, it is not practical to lay drip tubing to the root of every sapota plant. Maybe you may feed some extra water to coconut. The irrigation cycle itself would be different in different seasons. For example during summer the interval would have to be water every week, whereas in winter it could be water every three weeks. For pumping requirement one should consider the shortest cycle or interval. I do not have data about requirement of water for coconut. I assumed 4 inches every fifteen days to illustrate the method of calculation. To proceed with calculation, for feeding 1620 cubic meters per day, by running the pump for 8 hours every day, the flow-rate required of the pump would be 1620/8 = 202.5 cubic meters per hour = 2700 liters per minute = 45 liters per second. If drip irrigation would need only one-fourth, it would be 11.25 liters per second. It is logical to use drip irrigation for coconut, because once planted, coconut would grow for years unending. With most other crops like jowar, rice, wheat, sugarcane, etc., the field has to be ploughed and tilled every season. Drip tubing would also get uprooted and will have to be re-laid. This would be a problem even with sprinkler irrigation. This, I guess is the reason, why drip or sprinkler irrigation does not become feasible for many crops, though requirement of water is very economical for these methods of irrigation.
Please note that all the calculation is done assuming all 60 acres being at one place. But it is not so, as you have mentioned. It would be better to do separate calculation for the peice of land near the river and for the one 2 km away. Actually you plan to put water from the river into the wells in the field away from the river. Those wells will have some yield or rate of percolation of their own. If the total yield of the wells is 5 liters per second, then for drip irrigation you would need pumping only 6.25 liters from the river, even if all 60 acres are away from the river.
Anyway, having illustrated calculation to find the discharge for the pump, on to finding the head. You have mentioned maximum 37 feet height above the land near the river and 15 feet depth for drawing water from the river. That makes total height 52 feet = 15.85, say 16 meters. If 6.25 liters per second are to be pumped across 2 km distance, assuming farthest well is at 2 km, the frictional head in a 2.5 inch GI pipe would be 5.7 meters for every 100 meters, so for 2 km, total 114 meters. Instead if one would use 3" pipe it would be only 2.1 meters for every 100 meters, hence 42 meters and for 4" pipe at 0.7 meters for every 100 meters only 14 meters. Adding the frictional head to the height of lifting, 16 meters, the total head for the pump would be 130 meters for 2.5" pipe, 56 meters for 3" pipe and only 30 meters for 4" pipe.
You have also enquired about the type of pipe to be used, whether GI, PVC or HDPE? The issues involved are durability of the pipe, smoothness of the pipe, size or actual inside diameter of pipe and cost of pipe. In case of PVC or HDPE pipes, actual inside diameter is more important than just the size of pipe. Both PVC and HDPE pipes will be smoother than GI, but will have smaller actual inside diameter. Smoothness reduces the frictional head, but less inside diameter increases the frictional head. You would have noticed above that for 3" inside diameter the frictional head is only 2.1 meters for every 100 meters, but increases to 5.7 meters for every 100 meters for 2.5" inside diameter. In respect of durability, PVC is least durable, especially if it is exposed to ultra-violet rays from the sun. There is also the menace of the rats and rodents and you have to protect 2 km long pipe from them. For putting the pipe underground you would have to dig a trench. All that labour would not be needed for taying over-ground. From connsiderations of durability, safety and maintainability, cost of installation including cost of making trench, GI pipe may still be an option to keep in mind.
Coming to motor HP, to pump 6.25 liters per second against 56 meters total head for 3" GI pipe, assuming pump efficiency of 50 percent, one would need 6.25*56/75/0.5 = 9.3 HP, say 10 HP. Pump for same 6.25 liters per second but against total head of 130 meters for 2.5" pipe is likely to be less efficient say pump-efficiency of only 40%. Then motor HP will be 6.25*130/75/0.4 = 27.03 HP, say 30 HP. In fact if you can put pumps in series, at some distance, say at every 600 meters along the pipeline, with 3" pipe the head per pump can be only 19m. Each pump can be more efficient, say, 60%, then motor HP of each pump would be 6.25*19/75/0.6=2.64, say 3 HP. So, one can manage the situation with 3 pumps of only 3 HP each than one pump of 10 HP. And you will appreciate that 3 HP pumps are more commonly available, better mass-produced and hence more competitively priced than one 10 HP pump. Getting spares and repairing and maintenance of 3 HP pumps will be cheaper and more easily manageable than for 10 HP pump. With 4" pipe, requiring total head of only 30 meters, the motor HP required assuming only 50% efficiency would be 6.25*30/75/0.5 = 5 HP. Well, only one 5 HP pump, but 2 km of 4" pipe!
Now there is also the cost of power. Even if this may not be a significant cost to a farmer, the nation bears the cost, say, some Rs. 2 per unit of power, as the cost of generating the power and transmitting it to the ultimate user. For irrigating coconut you will run the pump for 9 months of the year, leaving out 3 months of monsoon. This means using the power at 8 hours per day, for 270 days i.e. for 2160 hours per year. If efficiencies of 3 HP (2.2 kW), 5 HP (3.7 kW) and 10 HP (7.5 kW) motors are taken as 80%, 85% and 90% respectively, for three, one and one number of 3, 5 and 10 HP pumps respectively, the energy consumed comes out to 3*2.2/0.8*2760 = 22,770 units for 3 Nos. 3 HP pumps, 3.7/0.85*2760 = 12014 units for one, 5 HP pump and 7.5/0.9/*2760 = 23000 units. So ANNUAL cost to the nation is Rs. 45540, Rs. 24028 and Rs. 46000 respectively. If the coconut field will be bearing fruit for 45 years after initial gestation of 5 years, irrigating the field for 50 years would cost the nation Rs. 11.385 lakhs, Rs. 6 lakhs and Rs. 11.5 lakhs respectively. The discussion above has brought forth different cost involved, cost of pump, cost of pipe, cost of laying the pipe, periodic cost of spares, repairs and maintenance and the cost of power. Even if a farmer would want to leave out the cost of power, the total of other costs calculated over an average period of 15 years, it would be a good idea to work out the Life Cycle Cost for different options and then make the most economical decision.
Q.    What factors will get affected if pipe size of 50mm suction & delivery is used for a pump having flange sizes 100mm by 100mm. Will it harm the pump in any way or will only the discharge get affected?
Frictional loss in 50 mm pipe will be 32 times higher than that in 100 mm pipe! This is wastage of energy. Cost of installation will be cheaper with 50mm pipe. But wastage of energy will happen for the life of the system. Also, if a pump is pumping with a suction lift of, say 6m, and ence effective pipe length of say, 8m and required flow-rate is 6 1/s, the friction loss in 50mm pipe will be 1.5m as against only 0.05m for 100 mm pipe. With further losses in strainer and footvalve and bend and eccentric reducer, etc. the pump may suffer cavitation.

FIELD PROBLEMS
Q.    We are regular reader of your Pump magazine. We are encouraged to see your response for different problems related to pumps faced in industries and suggestions sought from you. We are referring one of our specific pump related problem. We will like to have your comments on it. We have a multi stage steam condensate pump in our Urea plant which takes suction from tank at atmospheric pressure and discharges it at 45 Kg/CM2. This pressurised condensate is used at different locations in the plant for different purposes . Now due to some change in process conditions we require condensate pressure up to 32 kg/cm2 only. As pump is discharging at 45 kg/cm2 it is unnecessary loss of electrical energy. We are thinking to remove 2 or 3 impellers to reduce the discharge pressure.
Please tell us whether this is possible or not? What other problems we may face?
What is the general formulae to get the relation between number of stages and pressure delivered a pump?
Pump details are as follows:
Stages : 8 Impeller dia - 180 mm (same for all) Impellers are in series Balance drum provided at discharge end with a balance line d/s of this drum to suction line, Speed: 3000 rpm Intermediate bearing sleeve provided. Capacity : 20 m3/Hr
General relationship between pressure and number of stages is of direct proportion. You have 8 stages for 45 bar. Hence number of stages for 32 bar = 32*8/45=5.69 say 6. So, you can remove 2 stages! With 6 stages you would get 6/8*45=33.75 bar pressure. Power saving will also be in direct proportion. Assuming 60% pump efficiency, at 45 bar pressure and 20 m3/h the input power to pump is 20*45*10.336/270/0.6=57.25 HP. If you have a 75 HP motor, presently it is loaded to 76 oad. At 34 bar for 20 m3/h with same 60% efficiency, the loading of the motor will be 20*34*10.336/270/0.6=43.4 HP only 57.8%. Efficiency of 75 HP motor at such part load operation will be less. To get full benefit of reducing number of stages, you may use a 50 HP motor. You may check whether pump efficiency given in pump characteristics from manufacturer is 60% and accordingly whether using 50 HP motor will be appropriate. Actually manufacturer might have given you pump characteristics with different number of stages also.
Q.    The axial flow pump of capacity 7700m3/hr it installed in CLOSED LOOP system under 100 mbar abs vacuum in our plant. When pump was first trial run we observed severe cracking sound in the pump and thought for the possible cavitation due to starvation of pump. The pump was than stopped, acid drained, and on inspection some rubber pieces were found at pump discharge end. Thus, it was thought that these rubber piece were probably causing the abnormal sound. After having carried out the rubber repair of the vessel the pump was re-started Before starting the pump we ensured that the liquid level is up to the desired operating level and Initially the sound was low and intermittent but when vacuum was pulled up to 100 mbar abs. the sound increased abnormally and became continuous. Pump was stopped immediately we took second trial run but, the observations are same i.e., heavy cracking & metallic sound in the pump. We have now stopped the pump. We feel there is some design problem with the pump & its application.(NPSH ?)
Axial flow pumps are essentially high specific speed pumps. And high specific speed pumps would have high NPSHr. No designer can design an axial flow pump with low NPSHr. Vacuum at suction is only making the matter worse. Heavy crackling metallic sound is typical of a pump suffering from cavitation. Since there cannot be a design with low NPSHr, only option is to explore how NPSH available can be improved, so as to have positive margin over NPSHr. Options to be explored seem to be as follows.
1)     Whether suction vessel can be at higher elevation above the pump.
2)     Whether pump can be lowered to a level below the suction vessel
3)     If there is a throttling valve at suction to adjust vacuum at suction, whether the valve can be eliminated or whether the friction loss across the valve can be reduced by using a ball valve and preferably of higher size.
4)     Whether the piping on the suction side with all appurtenances like bends, tees, valves can be of higher size
5)     Whether vacuum can be less stringent
6)     If liquid being handled is at temperature higher than ambient, whether a heat exchanger can be added to cool the liquid to bring down the temperature. Since the pump is in a closed loop, rise in temperature of the liquid circulating in a closed loop is a natural phenomenon. Your observation that the pump starts off okay but noise aggravates after some running is indicative of rise in temperature causing vapour pressure of the liquid going up, thereby promoting cavitation. A heat exchanger to cool the liquid before it gets back into the suction vessel should then help.
Q.    In our system we require chilled water at 1000lit/hr and 40-60psig pressure at 7.0deg centigrade. Original pump was centrifugal type with open type straight vane impeller, 2900rpm, single stage pump Due to some reason we had to relocate the pump and chiller and piping length was increased by 30mtr but elevation of pipe was same but when we opened the pump for inspection we found its impeller vane(03 nos) got broken also impeller to casing clearance was 2.0mm so we changed the impeller with new one and dia of new impeller was 1.0mm higher then old one but during functioning pump discharge pressure has dropped to 45psig and flow rate 350lit/hr also rubbing marks were found on impeller vanes so we increased the impeller to suction casing clearance by 0.5mm and pump flow has increased to 450 lit/hr but pressure further dropped to 25psig due to which we can't get the cooling system working properly we also shifted the pump near to system by increasing its suction line and reducing its discharge line since suction water tank of chiller is at higher level then pump center line but condition is same kindly tell me the correct method to increase the discharge pressure and flow of the pump.
For the small discharge and relatively high pressure, the pump ought to be a regenerative turbine type centrifugal pump (IS-8472). You have also mentioned straight vanes, which is typical of these pumps. The performance of these pumps is very sensitive to clearance between casing and impeller. It should be minimum. However you seem to have increased the clearance. Even 2 mm clearance you noted when you opened the pump was excessive. These pumps also have poor efficiency. The commonplace 0.5 HP mini monobloc pumps are of similar type, i.e. regenerative turbine type. These pumps are readily available. So, it would be a ready solution to replace the pump than to repair or rectify such tiny pump. Another option is to use a multistage pump. Pumps made in stainless steel sheet metal and in compact vertical version are becoming commonplace. Such pumps would give much more flow than 1000 lph, you are looking for. But if more flow is acceptable, you may not worry about the excess flow. Otherwise you may have to bypass the excess flow back to suction. Even with bypassing the excess flow, the pump may prove more energy-efficient than the regenerative type.
Q.    We have 250 meters of pipe(1" sch40) existing for pumping water to our system and we have got some problem in our existing pump (centrifugal type, single stage, 2970 rpm). We have other spare pump of same rating and motor HP but its suction is of 32mm so my query is can we put expander in high speed centrifugal pump's suction (i.e 25mm X 32mm) because i have seen reducer in the suction side and expander in discharge side only. Will this effect in pump flow rate, discharge pressure, power consumption and cavitation in impeller. Or we will have to change whole piping or pump.
For trouble-free performance of pump, one should be concerned of friction losses on suction side. If present pump had 25 mm piping on suction side and has worked okay until it developed trouble, using a new pump with 32mm suction size along with an 25mmx32mm expander between 25mm pipe and 32mm pump suction should not give any new problem, since the expander would not add to the friction losses in suction. But better to have a long expander and fit it without increasing total length of suction piping.
Q.    Collecting Data on Water Market. For the Academic Purpose I have to identify
1.     Identify Market size
2.     Identify Applications
3.     Identify Consultants / Contractors.
Can you please send any of the old presentation on Water Market if you are having ready made?
Water market for pumps is too broad a subject. Following list identifies 26 situations where pumping water is inherent!
1)     Large schemes for transmission of water from source to distant settlements. For example Mumbai gets water from Powai, Vihar, Tansa, Vaitarna lakes and from Bhatsa dam, which is about 60 km from Mumbai. Sardar Sarovar Scheme plans to take Narmada water to water-starved region of Saurashtra, And at National level special cell is set up to explore Interlinkage of rivers to benefit water-scarcity areas from water-surplus resources.
2)     Domestic water supply. All municipal water supply systems guarantee water only into ground level reservoir. Pumping to households has to be managed by societies. Lakhs of mini monobloc pumps help people to steal water from municipal mains!
3)     All water purification systems, whether of alum-dosing, sedimentation, flocculation, filtration, reverse osmosis etc. need pumps
4)     Swimming pools, amusement parks, decorative fountains
5)     Flood management
6)     Fire fighting
7)     Cooling water circulation in process plants
8)     Boiler feed
9)     Process water e.g. brines in paper mills
l0)     Effluent treatment and disposal systems
ll)     Sewage and waste water systems
12)     Storm water or cellar drainage
13)     Hydro-electric power generation or pumped storage power generation
14)     Dewatering at construction sites
15)     Dewatering in mines
16)     Fluidised transport of ores
l7)     Coal washeries
18)     Dredging of river basins, canals, sea shores
19)     Minor irrigation i.e. agricultural pumping
20)     Lift irrigation or medium size irrigation
2l)     Marine vessels, e.g. bilge water
22)     Aquaculture ponds
23)     Injection water for improving yieid, of oil wells
24)     Heavy water in nuclear power plants
25)     Water for high pressure jets for cleaning, gouging, cutting
26)     Water in HVAC systems.

Tech Quest

We would like to increase the capacity of a two stage pump by increasing impeller diameter.

Can you please suggest the procedure for the same? Does affinity low apply to individual stages of multistage pumps also?

A two-stage pump is actually two pumps in series, built in unified construction. So, individual stages can very much be handled. But you have to first handle first stage first. Because it is the discharge of the first stage which will move forward to the second stage. The second stage cannot deliver any more flow, than what it gets from the first stage.

One of our KSB multi stage pump failed due to trust bearing (tilting pod type) wipeout at normal side. Please let me know the reason for the failure. This bearing provided with pressure lubrication. This pump is having balancing piston arrangement. Then how calculate the balancing pressure for the particular pump. This pump used for naptha circulation with suction pressure of 6.5 Ksc and discharge pressure of 52 Ksc with temperature of 200 degree C at the speed of 3000 RPM. This pump having fast coupling and turbine driven with continuous service. Kindly suggest to rectify the above problem.

Tilting pad thrust bearing in a pump suggests that it ought to be a vertical pump. But there is also mention about balancing piston. Once one provides the tilting pad thrust bearing, one may not complicate the construction with balancing piston. But it seems, this particular pump does have both the arrangements together.

A bearing fail we would happen primarily and most commonly due to failure of lubrication. And it is mentioned that it is a forced lubrication. If there is a choke-up in the lubricant feed line, the gauge of the lubrication system will show pressure. But oil may not be reaching the bearings. The choke-up in feed line may even cause rupture of the line, then oil not reaching the bearings at all.

Balancing piston arrangement is also susceptible to jamming due to failure of lubrication. If there is a lubricant feed also for the balance piston, and if the balance piston is upstream of the bearing as far as the flow of the lubricant is concerned, jamming of the balance piston can cause starvation of the thrust bearing for lubrication. In vertical pumps, misalignment can be the root cause for many of the problems. Misalignment may itself set in due to vibrations. For Root Cause Failure Analysis, one may have to then investigate the cause for undue vibrations.

I have been regular reader of your informative magazine Pumps India. I am on deputation, from IFFCO side, as Plant Shift Manager in Oman India Fertiliser Company (OMIFCO) Omon, It is a joint venture of fertilizer company IFFCO and KRIBHCO from India and LNG from Oman.

Plant has two stream of Ammonia (1750 MT each/day) and Urea (2550 MT each/day) capacity. It has been commissioned few months back and is running trouble free since then.

I have one specific query with reference to high pressure multistage centrifugal pumps installed in Urea plant and need your view on it.

In Urea plant we have Ebara Japan make high pressure multistage centrifugal Carbamate pump taking suction from a vessel operating at 17 ata. Pump consists of a booster pump (single stage centrifugal pump) boosting pressure from 17 to 22 bar and main pump (six stage centrifugal multistage bock to back 3 + 3) pressurizing liquid from 22 to 156 bsr. Booster as well as main pump is driven by a single Motor. Booster pump is directly coupled with motor at one shaft end while main pump is coupled at other end with on in between gear box to increase speed.

Pump discharge line has a recycle control valve. This is for initial start up and to maintain minimum flow through the pump. Down stream of the tapping for recycle valve is a heat exchanger called pre heater where discharge fluid is heated by hot condensate. Downstream of this heater is a control valve to maintain the carbamate flow to system as per the process requirement. Pre heater raises the temperature of fluid form 80 to 105°C. Carbamate solution is mainly solution of ammonia and carbondioxide in water. Vapor pressure increases as we increase the temperature.

Similar configuration we have for Ammonia high pressure multistage centrifugal pump (Ebara make) in urea plant. Here liquid ammonia is pressurized from 22 to 240 bar and heated in a heat exchanger form 36 to 94 0C before being fed to system.

Fluid characteristic (as per material balance) before and after preheater in case Carbamate pump is as under.

            Before    After

Temperature, 0C      80     105

Pressure , bar abs      147     147

Density, kg/M3          920     880

Volumetric flowrate m3/Hr  90     95.81

Piping Dia Inches        6      6

In actual practice pump discharge pressure is nearly 180 bar and to maintain flow through pump recycle valve remains always opened.

Is there any formulae to know the minimum load at which a pump can run or it varies from pump to pump and is given by pump manufacturer?

Please tell us what will be difference in behavior of pump (pump discharge pressure and ampere) when the pre heater is in line and when pre heater is out of line.

The behaviour of a pump in two different situations of a given system can be derived by plotting the system characteristics for the two different situations and seeing the intersection of the system characteristics with the pump-characteristics. In the system with preheater in line, the flow after preheater is 95.81 m3/h as against 90 m3/h before preheater. Pressure is same i.e. 147 bar. So, it is not clear, where the extra flow comes from. Against a given pressure of 147 bar a given pump cannot give two different values of discharge, 95.81 m3/h and 90 m3/h.

Is there any formulae to know the minimum load at which a pump can run or it varies from pump to pump and is given by pump manufacturer?

Minimum safe flow has to be obtained from pump's data. It should not be attempted to work it from any formula. Most formulae rely on empirical values of co-efficients etc., which may not stay valid for a given pump in a given situation. Basic anxiety of minimum safe flow is related to temperature rise and corresponding rise in vapour pressure and consequent possibility of cavitation and thermal stresses to pump components. People have tried to figure out formulae for temperature rise. But I would recommend going by manufacturers recommendations for minimum safe flow. Minimum safe flow has other implications of increase in radial and axial thursts, instability in pump's H-Q characteristics etc, all of which are not addressed by only the temperature-rise formulae.

I am a Mechanical Engineer working in Aditya Birla Nuvo Ltd, Veraval, Gujarat as maintenance engineer.

I want the information regarding the screw pump, multistage centrifugal pump and multistage boiler feed pump my questions are :

1 How I can do audit of Split Casing/normal centrifugal pump on the working site?

2.what is easy and practical method for calculation of efficiency, flow rate, NPSH? Please provide with Basics!

3. How the actual Designing Of pumps is Carried out (Provide in steps)? What Parameters are needed in any design?

1 How I can do audit of Split Casing/normal centrifugal pump on the working site?

Some points to be checked are -

1a) Can ratings for head and discharge for the given site conditions be reduced?

1b) Is the rated duty near to the Best Efficiency point of the pump?

1c) Is the efficiency of the existing pump the best for the rated duty? One needs to know the "Bench-mark" values of pump-efficiencies from International data.

1d) By Life Cycle Cost analysis, will replacement of the existing pump with one of higher efficiency prove cost-effective?

2.what is easy and practical methods for calculation of efficiency, flow rate, NPSH? pl. provide with Basics!

"Basics" should be studied from text books. This forum is not a training program.

3. How the actual Designing Of pumps is Carried out (Provide steps)? What Parameters Are needed in any design?

Basic parameters for starting off with Pump design are Q,H and rpm, materials of construction and viscosity (if relevant). Different authors have outlined different approaches with different empirical equations and assumptions.

We have a 5-ton capacity Boiler. For this we are using multistage Grundfos make pump. Mainly problem which we are facing is continues passing of disk type NRV (non return valve) due to passing of NRV & back pressure pump get damaged. While taking the leek test and hydro test NRV operation found normal beside this sometimes NRV passes during operation . NRV position is vertical, & exactly below the long radius bend pipe which enters at the boiler. Please suggest us what will be the reason of passing of NRV and give me the solution

NRV should be near to the pump. NRV being near to the boiler, "just near to the long radius bend" leading to the boiler, pump is getting all the head in the vertical pipe up to the NRV. Whenever the pump is stopped, the head in the vertical pipe is causing the back pressure on the pump. So, because of its position too much above from the pump, NRV is not able to do its function of stopping the back pressure. The differential pressure on the NRV is also less to that extent and the seating is not having adequate differential pressure.

A pump used for Hot oil circulation (sp.gvt=0.7). Its head is 48 mtrs and flow is 73 m3/hr. How to calculate the motor kw required to drive the pump? Please send the formula for calculation of motor kw required to drive the pump.

For non-viscous liquids, the procedure is to first find Input Power (P) required by the pump. Formula for this is

P in kW - (sp.gr.)*(Q in l/s)*(H in m)/102/(Pump Effy)
.......Eq. 1

Here Pump Efficiency is better known from curves of manufacturers.

In the absence of curves of manufacturers, it can be estimated by using a formula as follows -

Effy = 0.94 - 1/(13.2*(Q in l/s))^0.32-0.29*(0.32-log(0.047*Ns))^2.......Eq.2

In this formula, Ns is specific speed, i.e.

Ns=n*(Q in m3/s)/H^0.75.......Eq.3

Note, Q in m3/s i.e.(Q in l/s)/1000.

For 'n' one has to know the desired or usual speed of operation of the pump.

Value of efficiency obtained from Eq. 2 will be the value of efficiency, if the desired duty is at the best efficiency of the pump with best design and with maximum diameter of the impeller. Since such combination cannot likely be at good probability, one should modify Eq. 1 as

P in kW = (sp.gr.)*(Q in l/s)*(H in m)/102/(Pump Effy)/0.8.......Eq.4

One should further provide a margin over the value obtained from Eq. 4, towards likely variation in system characteristics. The user would know this best.

One good aspect of this procedure thru' four equations is that it also checks up whether Ns will be in advisable range of centrifugal pump design. Advisably Ns>10. If a 4-pole motor was to be preferred, then the speed would be around 1450 rpm.

In milk industries pump are necessary my question is the construction of the pumps is food grade or not? If it is food grade what grade it is?

Dairy pumps are usually made in stainless steel. Apart from the MOC, important constructional feature of dairy pumps is the finish of the hydraulic passages. Dairy pumps are usually made by fabrication from SS plate material. The idea is that milk should not stay stuck in the surface roughness and it should not have chance to become curds before the second batch of milk is pumped.

The pumps are given a rinsing wash at the end of every batch and the pumps are periodically opened and cleaned. To facilitate this the pump casing if often made with a hinged clamping, so that the pump can be opened as easily as opening the door of our house at the arrival of a guest. The impellers are also usually open impellers, not even semi-open impellers.

The piping connections are also clamped type with the pipe-end having plain faces of two pipes butting against each other and the clamp riding over tapered faces on each pipe end. Such connections are also popular in SS piping used extensively in pharmaceutical industry.