Frequently Asked Question-30
- By Jayesh Patel
- Published 01/31/2010
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.
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.
