Pump design rate vs. Pump Amps 1

[Falcon03]

Hi

What are the relationship between the pump Amps and the rated flow rate?

We are maximizing the feed throughput to our Crude unit; we are running with two booster pumps out of three. Each one has the capacity of 27336 m3/D (design). While based on our current condition (actual) we are running with two pumps as I mentioned above at the rate of 60500 m3/D. Therefore, based on our data sheet we are running at over design condition if we talk about feed capacity, whereas, the actual Amps is lower than the design.

I would like to make sure that we are running under safe condition and to make sure that there is no need to run the third pump. Thanks for your help in this matter.

 

[BobPE]

Falcon03:

I would tend to believe your data sheets on the pumps and say that you are running them over capacity. Whether or not that third pump will help, we can't tell, you didn't give us enough information. The relationship between pumps and power depends a great deal on the type of pump you are using. Often times the electrical system is designed with greater capacity than the requirements for the pumps and shouldnt be used as a basis for design worthiness of the pumps.

BobPE

 

[electricpete]

The motor current should roughly follow BHP.

BHP vs flow is in general not a monotonic increasing curve and depends on pump type as Bob says.

The shape of the BHP vs flow curve for various types of pumps was discussed at:
thread407-27055

If a single-stage axial flow in general you will draw less current as you increase flow. Many mixed flow multi-stage pumps have a peak in BHP vs flow near BEP, and will decrease above that.

If you are that part of the curve than there is no concern for your motor. But there may be concern for operating near runout condition. Does it cause excessive vibration, does it cause cavitation etc.

 

[Falcon03]

Thanks guys,

To answer your question electicpete: No we have not noticed any problems with our pumps with the current operation windows. No vibration and cavitations. I just checked the AMP meter in our SUB it shows 210 amp vs 254 design. I will try to find more information and will get back to you. Thanks again

 

[vanstoja]

If the flow is equally split between the two running pumps,then they're both operating at 110.7% of design flowrate which is well below the 120 to 130% maximum flow generally recommended for centrifugal pumps that do not have a specified maximum flowrate (usually presented as a minimum system resistance curve for a plant that has varying, unthrottled flowrates). Reduced current draw above design flowrate just means that the peak current is closer to the design flowrate which may be at or near the design best efficiency point (bep) flow. The biggest concerns for very high flowrates are poor incidence angle-induced impeller channel backflow (recirculation) instabilities and, for fluid film unidirectional, toward-impeller loaded fluid film thrust bearings, the possibility of thrust "crossover" to toward-motor unloading which causes shoe flutter damage in Kingsbury-type thrust bearings.Dual-directional Kingsbury thrust bearings always have one set of shoes unloaded and prone to flutter damage and so are not a recommended design feature.

 

[electricpete]

vanstoja - I know you have talked at length before about the complex causes of thrust. It's hard for me to imagine decreasing thrust with increasing flow. I always pictured the change in momentum of the fluid was the biggest component of the thrust.

 

[electricpete]

... I guess I forgot about differential pressure across impeller which works in the opposite way. (increasing flow corresponds to reduced pump dp as we move along curve).

 

[BobPE]

vanstoja:

they make a lot of pumps without specified flow rates!!!! I see them in peices all the time and they dont work well....LOL Seriously, consider the original question, they are trying to increase capacity and are sacrificing the pumps as usual to accomplish this. The only way I can see increasing the flow like that is knowing you need more flow and changing the system curve with two pumps runing to get it. You won't hear signs of cavitation too easily in crude, its a great masking fluid for sound. The 15 to 20 percent range of operation is for the one pump running, with two in service, this range no longer holds. You would have to look at the combined pump curve with his system curve to see where he stands with the prolonged survival of the pumps.




BobPE

 

[d23]

Falcon03:

27336 M3/D is over 173000 BFPD or 5000 GPM. You have big pumps!!!

I would like to add few things to all of the help offered so far that you may want to consider:

1)
Pump designs include calculations for both the fluid gradient and fluid viscosity. If your actual operating conditions differ from the design data it will cause errors in the predicted operating conditions. As an example if you’re pumping a low gravity crude the gradient was assumed at a given temperature. If your actual fluid temp is different than the design considerations both the gradient and fluid viscosity will change therefore the actual assumed operation will change.

1A)
If you do have a low gravity crude you may also have entrained gas trapped in the fluid that will add to a change in the fluid gravity. If there is gas it may be breaking out or expanding in your flow line after the pump generating lift that was not accounted for in the original design. This would be a facility separator or chemical type thing that you should consider. Personally I have had sub-pumps in wells with both low gravity oil and a low GOR that required gas separation in the wells to prevent gas in the pipe or sales line. This phenomena occurred after both a heater treater and storage tank. Makes me wonder whose gas correlations are correct!

2)
If you are running on a VFD or have power factor correcting capacitors you may actually be operating at the predicted kW load, but the amps you're reading could be deceiving you. It is my understanding that on the input of a VSD, where you have a full wave rectifier the power factor will appear as unity or 1.00. This would also be the case with well designed power factor correcting capacitors too. It then depends on where you are actually reading your amps at as to what you should expect to see. I'm sure that someone like electricpete could explain this better than me, but the fact is that it takes about 746 kW per HP for an electric motor and kW may be a more accurate way for you to consider your problem.

3)
If your system was designed for all three pumps to operate at the same time and you are only running two the pressure or head will be lower causing a higher flow at this time. You may want to compare both your current system operating curve and your original system design curve.

4)
I have to believe that you have some big pumps and high HP motors. If there is an actual failure you will expect the manufacture to at least consider warranties. I would recommend that you go to the manufacture now and get their input, in writing, on your current operation. This may save you headaches down the road.

4B)
Before you contact your pump manufacture you need to get the fluid temp at the pump intake and pump discharge as well as a produced fluid chemical analysis. You have a crude oil that should have very good lubricating properties, but some minor chemical additives like amine, polymers etc. require self lubricating bearings made of materials like graphite.

4C)
Another down side to your good lubrication is the specific heat of the oil. Oil by comparison to water does not transfer heat very well. By changing your operating or efficiency point you may have heat/pump seal problems that needs to be addressed reasonably quick.

5)
During the design phase of your project your pump supplier may have believed they know more about your operation than you do. They may have coated the stages of your pump with something that you're not aware of. As an example if they expect that you will have entrained H2S or CO2 they may have coated the stages with some form of coating to protect the staging metallurgy. If this coating has a thin, more smooth finish than that of the original metal your efficiency will improve or conversely your load amps will go down.

6) FYI
For the short term a data acquisition system may represent more of an expense than value, but after some run history by getting volts, amps, flow rates, pressures, temperatures etc. you may be able to change your pump replacements into pump repairs. Just food for thought!

The big seven)
I wish I would have sold these pumps!!!!

I hope this helps!!!
David

 

[Kawartha]

Falcon03,

Sorry to jump in so late on this thread, but thought that some basic pump info may be of help.

The first recommendation that I would make is that you obtain a copy of the pump performance curve. On this size of pump (> 5,000 gpm), you must have some witnessed, "Certified Performance Curves" available. If you don't have these, the pump supplier can provide copies.

Secondly, pump brakehorsepower is based on flow rate and differential head (not flow rate only), so depending on the pump specific speed and operating point on the curve, the BHP can drop with increasing flow rates.

BHP = Q (flow rate) x Hd (differential head) x Specific Gravity / 3960 x Hydraulic Efficiency.

As a general rule: low Specific Speed (Ns) pumps (radial flow pumps with Ns of less than 1,000) have continuously increasing BHP requirements with increasing flow rates. Medium Ns pumps (mixed flow pumps with Ns above 1,000 to 4,000)have increasing BHP requirements up to about bep, and then reducing BHP to runout. High Ns pumps (axial flow with Ns greater than 6,000) can be considered to have continously lower BHP requirements at higher flow rates.

If your pumps are relatively high specific speed then the lower bhp that you have observed makes sense.

 

[electricpete]

Kawartha - I am happy to see that your description of the general pattern of BHP vs flow curve matches what I have pieced together with assistance from folks here (as described in the link above). I have not used the concept of specific speed, but my observation was that:

-pure radial flow usually have continuously increasing BHP vs flow
-pure axial usually have continuously decreasing BHP vs flow
-mixed flow have more complex pattern which often peaks at BEP and decreases both directions from there.

The other part I mentioned was that the pure radial and pure axial flow seem to appear only in single-stage pumps... the mixed flow seems to occur in multi-stage pumps.

 

[Kawartha]

Electricpete,

Sorry, I didn't mean to denigrate your comments. Just thought that the thread should get back on track.

However, the full range of specific speeds apply equally to single and multistage pumps. Multi-stage axial flow pumps are not uncommon, and radial flow designs are extremely common on multistage pumps such as deepwell designs (high head / low flow). Likewise, most of the end suction (ANSI and API) pumps are mixed flow design.

Regards

 

[electricpete]

Hi Kawartha. I didn't view your comments as negative at all. I was just looking to bounce my understanding off of you. All of the multi-stage pumps I have encountered in my limited experience have been mixed flow. Thanks.

 

[vanstoja]

Electricpete,
Generally momentum thrust change is not governing for axial hydraulic thrust loading. It depends on the angle of the flow turn which is 90 degrees for pure radial (low specific speed) impellers and zero for axial flow impellers (high specific speed). The thrust curve change with flowrate tends to match the head-flow curve trend particularly on the high side of bep flow since it is the product of static head minus pressure losses therough the leakage path times the unbalanced front-to-back impeller area.

Bob,
I'm used to seeing pump designs operating around +/- 20% of design best efficiency flowrate because the powerplant system designers simply couldn't predict the targeted system resistance curves to any better accuracy...repeatedly on several plant designs in succession. Consequently, I'm likely to view a 10% off-design condition as not too disturbing. I'm used to pressurized closed systems with prescribed large cavitation prevention margins built in by conservative minimum suction pressure operating requirements. This does not hold up well if there is only a small cavitation margin at bep flowrate. Then the inlet

flow incidence angles on the blade leading edges can turn "recirculatin" secondary flows into vortex strings breeding cavitation vapor bubbles in their coreswhich is decidedly bad news as you suggest

 

[Kawartha]

vanstoja / BobPE

Your comments regarding recirculation cavitation for pumps operating outside of the recommended envelope are very helpful. However, they may not be applicable to Falcon03's requirement. Falcon03 advised the design flow rate and the actual flow rate conditions. I don't think that he provided the best efficiency flow rate for the pumps, so without this information we can not determine the recommended flow envelope for the pump.

Be very careful when establishing maximum/minimum flows relative to bep. A very good reference on this issue is ANSI/HI 9.6.3-1997 "Centrifugal and Vertical Pumps - for allowable operating region". The preferred operating range (POR) varies considerably based on pump specific speed, the pump suction energy, impeller tip speed, and design of the pump.

As a general rule, HI recommends a POR between 70% and 120%. Depending on max. impeller speed, POR can range from 25% to 120% of bep. For higher specific speed pumps (axial flow design), the POR is much narrower than for mixed or radial flow pumps. I have had experience with some low horsepower (3HP), multistage, radial flow pumps operating at zero (0) flow for years - without damage.

Design engineers should always obtain the POR recommended by the pump manufacturer, as warranty will obviously be impacted by operation of the pump outside of recommended range.

 

[Falcon03]

Thank you all of your help.

Kawartha, I have checked your equation above and I have applied but I got different results. I would like to make sure that the above equation is based on English unit not Metric. Because we do metric unit. Thanks again.

 

[Kawartha]

Falcon03,

U.S. Customary Units.
To calculate Pump BHP: BHP = Q x Hd x S.G. / 3960 x Eff.

BHP - in HP
Q - flow rate in usgpm
Hd - differential head in feet of liquid
S.G. - Specific Gravity (dimensionless)
Eff. - Pump efficiency at Q (from pump performance curve)

Metric Units.
Formula for Metric, I believe is as follows.
kW = Q x Hd x S.G./ 366 x Eff.

Q - cubic M / hour
Hd - M

Regards,

 

[BobPE]

Kawartha:

I think our intrepretation of the problem stems from information supplied by Falcone3. he stated that they are over capacity as indicated on the pump data sheets. From this, one can speculate the problems associated with this overcapacity.

I can't agree with you more though on POR. This is the most widely misunderstood and neglected range associated with a pump. In my opinion, it is so important that I have it stamped on the Pump plate in my specifications.

BobPE

 

[Kawartha]

BobPE

Thanks for your comments. I can understand the confusion on this issue. It would have been more helpful if Falcon03 had provided the pump bep, but he only provided the design and operating flow rate. My suspicion is that the pump is a relatively high specific speed pump, and that the design point is probably close to or slightly on the right of bep. This would help to explain the observed lower operating flow hp vs. design flow hp.

Your comments regarding possible recirculation problems is pertinent and Falcon03 should do further investigation into this potential problem. As you suggest they may be close to the maximum POR. Since a lot of the answers will be provided on the pump performance curve - I suggested that Falcon03 obtain a copy.

 

[Falcon03]

If some one is welling to get the pump data sheet I will be welling to send it over to you. What I need is the Fax No. or e-mail address so I can scan the data sheet to you.

Thank you again and sorry If I wasn't clear enough in my original question.

To be frank with you guys I didn't get so much help from our pumping machine people as you provided to me, since I'm Chemical eng.

I'm sure I will need you assistance in the future.

Bye.