15/20 years back if anybody purchases any product having ISI mark he was proud to say this item is ISI marked product. DO you think people have same spirit today, No. The immediate reaction to ISI mark is No guarantee for quality.
In our pumps line the conditions are not different. Because there are many manufacturers. Eveybody manufactures the same product. For example take 3Hp monoblock pumpset. But in different size 90 frame, 100 frame, 112 frame and 132 frame, in different design, different weight different discharges, different head ranges, different prices, different input current. For a single item why there are these differences. If there is a quality control method there should not be such vast variations in one product.
if we analyze the current input factor alone, we find that 3Hp monoblock pump set manufactured from top quality raw materials consumes about 3 units per hour. Whereas the pump set not manufactured from top quality raw materials takes 4 to 4.5 units per hour i.e. the national loss behind a pump set is 0.5 units per HP per hour. Several lacs of such pumpsets are in operation in India which consume more units. Both of types pumpsets bear ISI mark and we do not know at what price ISI allow this, the electricity board overlooks this, government keeps mum on such prime issues.
If ISI have a correct quality procedure there should not be a loss making point to nation Even if ISI have procedures they are never maintained properly, never followed regularly by the ISI people and there are cases even the ISI officers have not visited the manufacturing units years together or if visited they were managed to overlook such points, if we say the ISI department is responsible for a major part of energy loss in the country it will not be a mistake.
There are several lacs of electrical connections applications of poor famers which are pending in each state and day by day the shortage of electric supply is increasing. If we study a case of 1000 nos. 3 HP pumpsets consuming more current and if these 1000 pumpsets can be replaced with good quality pumpset the lectric board can run 1500 pumpsets with the same units of current which was used for 1000 pumpsets. I do not understand why the electricity authorities are not thinking on this technical point seriously and adopting proper methods to check and control the loss. In India it is very easy that somebody makes mistake and the punishment goes to others and electricity department is doing this simple methods of recovering money against the loss from other users. But how long it will be possible?
So why the ISI authorities, electricity board and the government all together make a uniform method to stop losses incurring by using inferior quality products. Here the policy of major political parties are also one of the reasons for national loss. We can say in this loss making procedure ISI is a partner, political parties are a supporter, Government is a over looker and the innocent public the suferer.
Will there be a system of educating or compelling the people to use power saving products which consume only the permitted units of current to save the nation from electricity losses and shortages.

Author Mr. S.G. Nair

One of the claimed advantages of the centrifugal pumps over positive displacement pumps is their ability to operate over a wide range of flow. Since a centrifugal pump operates at the intersection of a pump curve and a system curve, by varying the system curve the operating point of the pump is easily changed:

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.