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The “two-thirds” Rule for Locating Sensors to Control Variable Flow Systems | A Field Perspective on Engineering
A Field Perspective on Engineering
Engineering lessons from the field
The “two-thirds” Rule for Locating Sensors to Control Variable Flow Systems
Posted on November 4, 2007
Recently, I was out in the field scoping out an existing facility for retrocommissioning opportunities. One of the
operating engineers pointed out that the differential pressure
sensor controlling the chilled water distribution pumps was located so that it sensed the pressure across the mains
leaving the plant. He had heard that you could save energy by locating the sensor two thirds of the way between
the pumps and the most remote load, so he thought that relocating the sensor might represent an opportunity.
But, he wasn’t exactly sure why.
For one thing, what did “two thirds of the way to the most remote load” really mean; physical distance or feet of
pipe or something else? He also was wondering why “two thirds” instead of “three quarters” or “seven eights” or
“fifteen sixteenths”? Finally, he was curious about exactly how the point where the sensor measured pressure
impacted the energy consumption of the system.
He knew that both the pressure and flow a pump produced impacted the power it required, all as stated by the
equation for pump power.
But he wasn’t sure how the relationships applied in a working system. Finally, he wondered if there were other
things to consider. For instance, if the sensor was located at remote point in the system, what would happen if
someone isolated that portion of the system for service? And from a practical standpoint, it seemed like it would
be a lot harder to remember the location of a sensor at some remote point in a 500,000 square foot facility versus
a location that was in the immediate vicinity of the pumps it served.
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These are all good questions and come up a lot in the field when I start talking about using a remote pressure
sensor to optimize the energy consumption of a variable flow system, be it an air system or a water system like we
are discussing here. So, I thought the subject might make good fodder for a few blog posts.
As a starting point, lets consider a simple system like the one illustrated in the system diagram below.
Our discussion will focus on the distribution piping network served by Pump P1; i.e. the flow path from A through
B, C, D, E, F, and back to A. For the purposes of our discussion, I’ve made a few simplifying assumptions.
The reference pressure is the pressure established by the expansion tank at the suction of pump P1 and is
assumed to be 15 psi (6.5 ft.w.c.).
The loads served by the system are identical; specifically at full load, both AHU1 and AHU2 require 400 gpm
and a differential pressure of 20 ft.w.c. at the point where they connect.
The pumps, piping network, and air handling units are all at about the same elevation; thus the effects of
elevation on the pressure readings can be ignored.
Piping lengths will be discussed in terms of equivalent feet. In other words, when I say that the distance from
point A to point B is 200 equivalent feet of pipe, I’m saying that if I were to convert the resistance due to flow
of all of the fittings between point A and B to an equivalent length of straight pipe, and add it to the actual
length of straight pipe, it would be the same as 200 feet of straight pipe with no fittings in it. If there were no
fittings, then the distance would literally be 200 feet, but as the number of fittings increased, the physical
distance associated with 200 equivalent feet of straight pipe would be reduced.
One of the key concepts that you need to understand with regard to this topic is that the pressure required to
move water through a system is a function of the flow in the system. This means that if the control valves on
AHU1 and 2 are both closed, the pressure reading at the discharge of pump P1 would be identical to the reading at
the tee where the supply piping splits to serve the two AHUs, even though the tee is three hundred equivalent feet
away from the pump. (Remember, we have assumed that both points are at the same elevation, so the impact of
elevation on pressure can be ignored).
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If the control valves in the air handling units start to open, the pressure at the tee will start to drop relative to the
pressure at the pump. The magnitude of the drop will be a function of the flow rate. Common wisdom is that this
relationship is a “square law”.
In other words, if you reduce the flow by 50%, the pressure drop associated with the flow drops to 25% of what it
was (50% of 50%), assuming nothing changed in the system (no valves changed position, the piping was not
modified, etc.). Research has shown that for most real piping systems the exponent is more like 1.89 versus 2, but
out in the field, we probably couldn’t measure the difference as illustrated by the following graph.
So for our purposes, we can still think of it as the “square law” instead of the “one point eight nine law”.
Now, lets look at the pressures required by our system in two different operating modes. In the first operating
mode, the system is at full load. Both AHUs are using 400 gpm and the pump needs to deliver 20 ft.w.c. at the tee
serving the AHUs while moving 800 gpm through the mains to and from the air handling units. The total length
of the piping circuit is 600 equivalent feet; 300 equivalent feet from the pump discharge to the tee that serve the
units and 300 equivalent feet back to the pump suction.
In the second operating mode, one AHU has shut down but the other is still operating at full load. This means that
the pump still needs to deliver 20 ft.w.c. at the tee serving the unit, but only needs to move 400 gpm through the
piping network. The two operating modes are compared in the following table.
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Note that the friction rate for the second mode is significantly different from that associated with the first mode.
In the next post, we will look at the results in more detail.
David Sellers
Senior Engineer – Facility Dynamics Engineering
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The "two-thirds" Rule; Some Bottom Lines
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The "two-thirds" Rule for Locating Sensors
to Control Variable Flow Systems - Part 3
Retrocommissioning Findings: Make Up Air
Handling System Simultaneous Heating and
Cooling – Data Logging and Testing
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8 Responses to The “two-thirds” Rule for Locating Sensors to Control Variable Flow Systems
nadim chabab says:
February 25, 2013 at 12:59 am
Further your above valuable notes, please note that in our application 2 way constant flow pressure independent control
valves are installed. I beleive the theory of two-third still applicable. but I would like to conifrm what do we mean by two
third. If for example the pump pressure head is 3Bar, then the two third is 2 Bar, So the sensor should be installed in a
location in the longet run where the Differentail pressure is 2 Bar when the system is operating at full load???
Reply
David Sellers says:
February 27, 2013 at 11:34 am
Hi Nadim,
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I would agree with you that even though you have pressure independent control valves installed, the two-thirds
rule concept will likley be worthy application.
To be clear, the “two-thirds” in the two-thirds rule is a general reference to distance, not pressure. What it really is
trying to say is that you should put the sensor as far out in the system as possible. In a perfect world, we would
figure out which run has the highest pressure drop and then put the sensor at the end of it, as close to the load as
possible.
And in a simple system like the one in my example, you would likely do just that. Meaning you would install the
sensor to pick up the pressures at C and D in the diagram. But few systems are that simple, and all systems are
dynamic. So, its not out of the question that in a complex system, under some load conditions the critical branch
(the one with the most pressure drop) might move around some due to the dynamics of the system. Or, someone
might make a change to the system (add a load, change the setting on balancing device, etc.) that would impact
which branch was the critical branch.
So, to provide some measure of protection from problems like that, the two thirds rule is in effect, saying “don’t
put the sensor all the way at the end of the critical branch; rather locate two thirds of the way there so you have a
safety margin if something changes”. Two-thirds is an arbitrary number; a while back, I was talking with Chuck
Dorgan about it (sort of the grandfather of the commissioning industry in a way) and he said that ASHRAE had
done some research and nobody really knew where it came from, but it probalby came from a control vendor
engineering bulletin.
So bottom line, the rule could have been the “one-half” or the “five-eights” or the “three-quarters” or the “fifteensixteenths” rule, dependiung on how much safety margin you wanted. The idea is to put the sensor as far out the
pipe as possible so that the pressure you control at reflects the square law pressure drop due to flow that occurs
and minimizes the pumping energy you need use to deliver flow to your loads.
In a working system, frequently the best way to find this point is via some field testing. Specifically measure
pressures around the system at points where you can gain access (vents and drains are good opportunities for
this). Do this under a number of different operationg conditions and compare what you measure to what you think
you need at that point to deliver design flow. Most of the time, a pattern will emerge that will lead you to both the
appropriate location for your sensor and also the ideal set point to maintain.
Hopefully this clarifies things and answers your question, but if not, don’t be afraid to ask additional questions.
Best,
David
Reply
Rajulukose says:
February 5, 2014 at 3:31 am
Respected sir,3 floor bldg pump room roof 3rd floor bldg length 80mtr 50mtr width with 4nos riser how to calculate
diffrencial sencr for chilled water system.Raju.hvac supervisor .Bahrain
Reply
David Sellers says:
February 7, 2014 at 11:47 am
Hi Raju,
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Its been a busy week so not much time to devote to the blog lately. If I am understanding you correctly, you are
trying to figure out the correct set point and location for a differential pressure sensor in an existing chilled water
system. If that is the case, then the fast answer is that instead of doing a calculation, you can do a few tests and
take a few readings and let the building tell you the answer.
Specifically, you would start the process by going around and taking differential pressure measurements at the
various loads on a day when the system has a pretty good load on it (i.e. control valves are allowing a lot of flow).
Even if there are not gauge ports in the piping, usually, at least in the States, you can find a point to connect a
gauge, often in the form of a vent or drain connection. When you take the reading, it would be good to note the
position of the control valve if you can and also the pressure being produce back at the pumps if possible. The
latter can often be handled automatically by running a trend or using a data logger.
Once you have done that, when you compare the readings, you will probably see a pattern that points to the critical
load or loads. One of those loads is the probably the location for the sensor. If it is a complex system, them maybe
you put 2-3 sensors in at various points and let the control system pick the worst case.
In terms of the proper set point to maintain at those locations, the exact number will depend a bit on what is in the
pipe between the two points where you measured pressure. Say for instance you picked up the high pressure at an
air vent ahead of the control valve serving the coil and the low pressure at the drain on the leaving side of the coil.
That would mean that the pressure drop was generally composed of the control valve, the pipe in the circuit
between those two points, and the coil pressure drop.
Typically, the pressure drop in the piping is fairly insignificant in the context of the coil and the control valve. So,
you could get in the ball-park for the right number by looking up the coil and control valve performance data in
the submittal drawings and then adding a little bit more to that value to come up with your set point.
So in terms of the example above, you might look up the coil submittal or call the coil manufacturer and discover
that at design flow the coil pressure drop was 12 ft.w.c. (about 35 kPa I think; not sure what your units system is).
And, if you looked up the valve, you might discover that it had a flow coefficient that resulted in a wide open
pressure drop of 10 ft.w.c. (about 30 kPa if my little unit conversion program is doing its thing correctly). (If you
need to know how to use a valve flow coefficient to calculate the valve pressure drop from the flow, the MCC
Powers valve sizing bulletin I link to from the resources on the blog or the Honeywell Gray Manual I link to in one
of the posts has that information in it).
So, based on your research, you now know that you need at least 12 plus 10 or 22 ft.w.c. (66 kPa) of differential
pressure at the points where you measured it to deliver design flow to that particular load. If you add 1-2 feet to
that to cover the pipe and fittings in the circuit between your two measurement points, you probably have a pretty
good set point to start with.
Once you have the necessary sensors and control logic in place, you can use trend data to help you fine tune the set
point. Say for instance on a hot day, you discover that you are not quite holding the required supply temperature
from the coil, even though the valve is wide open and that back at the central plant, you still had reserve pumping
capacity available (meaning the pumps had not been driven to full speed). In that case, I might increase the set
point by 1-2 ft.w.c. (3-6 kPa) and keep watching what happened until I found something that worked on that
design day.
In contrast, if you noticed that the control valve to the load was always throttling a little bit, even on a heavy load,
then you might conclude that you could drop the set point 1-2 ft.w.c.. The point is you can use the building
operational data to fine tune the system.
So far, I have discussed this as if you were trying to figure it out for an existing system. But the reality is you can
use a similar technique for new system and then fine tune it via the commissioning process once it comes on line.
The only part of the technique that is different for new construction is that you either have to make an educated
guess regarding where the critical load(s) are or do the hydraulic calculations. Typically, unless a load has a very
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high pressure drop relative to the others, the critical loads will be the ones with the longest piping runs. So, you
often can decide where to locate the sensor with a bit of engineering judgment, especially if you provide a couple of
them at different locations.
I need to head off and do something else at this point, but hopefully this gives you enough to go on.
Thanks for visiting and supporting the blog and for asking a good question.
Best
David
Reply
sureshbob says:
August 1, 2016 at 3:51 am
HI This is Suresh,
I have one AHU which is having the two branches duct in different length. Now where I have to install the pressure sensor to
drive the VFD.
Note : In all Duct have the VAV’s and 1st Branch length is 30 meters 2nd Branch is having 15 meters.
Reply
David Sellers says:
August 16, 2016 at 3:56 pm
Hi Suresh,
Sorry for the slow reply; I have been pretty overwhelmed with stuff the past few months and have not even had a
chance to post anything.
As you probably suspect, the concept regarding where you would locate the sensor for an air handling system are
identical to the ones I discuss for a pumping system in the post you are replying too. To find the exact, right
location, you would have to do a lot of math of course. But generally speaking:
1. I would pick a point in the longer duct because that is more likely to be the duct run that set the fan system
static. So you might select 2/3 of the physical distance down that duct. Once the sensor is installed, you could fine
tune the set point there as necessary to ensure that you had adequate pressure everywhere else. Ideally, there
would be one terminal unit that was nearly wide open someplace in the system.
2. You might consider installing a sensor in both duct branches about 2/3 of the physical distance down the duct
and then using the software to select the sensor that required the highest fan speed to meet set point.
3. Either way, in an air handling system, unless it is a small system or a system, I actually prefer to control for the
pressure at the discharge of the fan and then reset that set point based on what is going on at the remote location.
The reason I do that is that if you try to simply control the system based on the remote sensor, you may find that
there is a significant lag between when the fan speed changes and when the at the remote location sees the change
and reacts to it. That is because to pressurize the duct, the fan has to move enough air through it and the holes in it
(which we call diffusers) to set up a new pressure gradient and that can take some time. In some ways, the fan is
trying to inflate a really leaky balloon, so the volume of air it has to add to change the pressure is related to much
more than just the volume of the duct. I discovered this the hard way by blowing up a duct and then realizing what
David St. Clair meant when he said in his book about loop tuning that “it is all about the lags”. (Here is a link to a
blog post that hooks you up with the book if you want https://av8rdas.wordpress.com/2010/12/13/resources-forhttps://av8rdas.wordpress.com/2007/11/04/the-two-thirds-rule-for-locatiing-sensors-to-control-variable-flow-systems/
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understanding-pid-control/). So, a loop running based on a signal picked up right at the fan discharge is fairly
immune to the transportation delay issue (along with a couple of other things). And using the remote sensor to
optimize its set point kind of lets you have the best of both worlds.
Hope this is helpful and again, sorry for taking so long to respond.
David
Reply
eyad says:
June 17, 2017 at 6:12 am
dear Sir,
what do you mean by two-third distance ? should i calculate the whole distance from pump discharge to the furthest point of
networks then to put sensor on 2/3 of it ? or to calculate load as tonnage then to take 2/3 of it ?
please could you clarify by example
thank you for patience .
Reply
David Sellers says:
July 2, 2017 at 11:34 am
Hi Eyad,
Sorry for the delayed response; things have been busy for me and I am behind on responding to comments.
The short answer to your question is that the two thirds rule is a distance based rule vs. a tonnage based rule. So
that means you would want to locate the sensor at the point that was two thirds of the way to the hydraulically
most remote load.
But, like most engineering decisions in something as dynamics as a building, that is not as easy to figure out as it
sounds. The further out the system that you go, the more energy you will save.
But, if it turned out that you had not picked the hydraulically most remote load after all, or it moved due to system
dynamics or changes in the use profile of the facility then if you were way out at the end of the system, the set
point you maintained there may not be the one you needed. So, the two thirds rule is a sort of compromise
targeted at maximizing the saving you achieve while minimizing the risk you are taking in terms of the location
and set point you pick and the system dynamics that can cause that to not always be the right place or number.
If you really were going for the most optimized sensor location, then you would identify the load on the system
that required the most pressure from the pump serving the system to deliver its design flow rate; i.e. the load that
was used to establish the design pump head. This load if often termed the most “hydraulically remote” load.
Usually, but not always, the hydraulically most remote load will be the load with the most feet of pipe between it
and the pump and we frequently assume this is the case when we do our pump head calculations.
But for large, complex variable flow systems, figuring out exactly which load is the hydraulically most remote load
can be a complex undertaking. In addition, as a result of system dynamics, the hydraulically most remote load can
move around.
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For instance, if the load that is determined to be the hydraulically most remote load on the design day at the
design hour with all loads in operation is off line and there is not flow to it, then it is no longer the hydraulically
most remote load. The critical load will shift to a different load that is in operation and requires the most pump
head to deliver its required flow rate to it at the current operating condition.
In a general sense, what the two-thirds rule is saying is that the further out the system you put the sensor
controlling the speed of the distribution pump, the more energy you will save. That is because when the sensor is
at the end of the system it “sees” the pressure drop due to flow that is required by the current operating condition
and automatically drives the pump to deliver that much pressure in addition to the pressure it is trying to
maintain at its installed location.
Since pressure drop due to flow varies as the square of the flow, you can capture the bulk of the benefit associated
with using a remote sensor. You can see this in the table that is provided at the end of the third post in the series.
Notice how locating the sensor at the two thirds point captures 77% of the possible savings compared to a base
case, which is a pump selected for best efficiency operating at the reduced flow condition associated with the
system at 50% load.
So, as I indicate in the last post in the series, the two thirds rule could have been the three quarters rule or the
fifteen sixteenths rule or the twenty-seven thirty-seconds rule; it’s simply a rule of thumb that tries to simplify a
complex engineering decision in a way to captures most of the potential savings by also protects you from putting
a sensor at a load that turns out not to be the hydraulically most remote load after all and having to deal with the
consequences of that.
As an aside, I should mention that Chuck Dorgan actually did a little research project for ASHRAE to try to figure
out exactly where the two thirds number came from. It turns out that, to the best he could tell, it came out of a
technical application bulletin one of the major control vendors had developed to guide their field technicians and
provide a easily remembered number that would save energy but keep things “safe” in terms of ensuring the
performance of the system.
Hope that helps; thanks for visiting the blog.
David
Reply
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A Field Perspective on Engineering
Engineering lessons from the field
The “two-thirds” Rule for Locating Sensors to Control Variable Flow Systems – Part 2
Posted on November 4, 2007
In my previous post, I looked at what happens to differential pressures at different points in this variable flow
pumping system as the flow rates drop off.
In the example, when one of the AHUs shut down while the other remained in operation at full load, the
differential pressure required at the pump dropped by about 40% relative to what was required at full load while
the flow dropped by by 50%. In the energy conservation game, our goal is to exploit this potential for reduced
power consumption while still meeting the requirements of the load.
Let’s revisit the pump power equation.
Both head and flow appear in the numerator. Clearly, if we could control the pump in a way that reflected the
reduction in both of these operating parameters, we would optimize its energy consumption.
Let’s take a minute to think through what would happen in our example on a design day when one of the AHUs
shuts down. For the purposes of discussion, I’m going to make a pump selection for our example using Bell and
Gossett’s ESP software.
Entering our design flow and head (800 gpm at 44.2 ft.w.c.) generates a number of selections. I chose the one
with the best efficiency, which comes at a cost premium of approximately 17% relative to the lowest cost option.
The 17% price premium buys a 2 % increase in pump efficiency. In economic terms, this may or may not be
justified based on the load profile.
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Holistically, one might argue that the price for improved efficiency is justified no matter what the load profile is
since the improvement likely represents a reduction in the use of non-renewable energy with a corresponding
reduction in emissions. (Of course, I’m from Oregon and grew up in the 60’s, so when I’m not busy hugging trees,
I’m probably out trying to save salmon; so you have to take all of that with a grain of salt).
If AHU1 shuts down while AHU2 remains in operation at full load, AHU1’s control valve closes. Eliminating one
of two parallel paths by closing AHU1’s control valve forces all of P1’sflow through AHU2. Increasing the flow
through AHU2requires more head that was being produced by P1 with both units online. As a result, the flow
through AHU2 will not double; rather the system will shift towards a new operating point/system curve with a
higher pump head than previously existed but at a reduced flow relative to what was provided with both units in
operation.
Initially, the reduced system flow is still in excess of what AHU2requires at design conditions. As a result, the
control valve serving AHU2 will throttle in response to the excess capacity. Ultimately, interactions between the
flow supplied by P1and the capacity the flow produces in AHU2 will cause the control valve to throttle the flow
through AHU2 to the design value of400 gpm. The figure below illustrates our pump curve along with the design
system curve associated with two units in operation and the new system curve associated with AHU2 operating
alone at its design capacity.
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Reducing Pump Power by
Pushing the Pump Up Its Curve via a Throttling Process
Back in “the olden days”, before variable speed drive technology had been perfected and made affordable we often
allowed the pumps to be pushed around on their performance curves by two way valves. While crude by today’s
standards, this approach was relatively simple and could save some energy as can be seen below.
By way of explanation and to provide some perspective, the “olden days is an expression my son Aaron would use
when he was younger to preface a question about something in my past. Back in the “olden days”, when I first
priced a VFD, it was for a 40 hp motor; the price was about $50,000 and the package as about the size of 2 motor
control center sections.
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On the plus side, when the pump was throttled:
The brake horsepower required at part load is reduced by about 2 bhp.
No special technology is required.
The theory of operation is simple and easy to understand.
On the minus side:
Pushing the pump up its curve moves it away from peak efficiency as can be seen by comparing the two
operating points on the pump curve above.
The head produced by the pump at part load is significantly above what is required , as illustrated in the
calculation below. Specifically, pushing the pump up its curve results in an operating point that produces 400
gpm at about 54 ft.w.c. as can be seen from the pump curve above. But, as can be seen from the calculation
with AHU1 off and AHU2 at full load, you only need 26.5 ft.w.c. of head at the pump to deliver 400 gpm to the
load and provide 20 ft.w.c. of head at the load. Initially, the extra head drives more than the required flow
through the load. This causes the temperature leaving the coil to drop below set point. In turn, the control
system closes the valve until the extra head is simply “chewed up” by the control valve and the design 400 gpm
flow rate is again achieved through the load.
The above design pumping head could lift valve plugs off of their seat if care is not exercised in selecting the
actuators. As a result, we may not achieve the desired reduction in flow and associated energy savings.
Plugging the desired flow and the head it would take to produce it into the pump power equation reveals that
in theory, we should be able to serve the reduced load condition with about 3.3 bhp.
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The bottom line is that while pushing the pump up its curve provided a simple way to reduce pumping capacity
and save energy, the potential exists to save even more if:
We can find a way to reduce pump flow and pump head at the same time.
We can make this shift in operating conditions while preserving the pump’s efficiency.
We can control the pump head and flow in a manner that provides exactly what the loads require, no more and
no less.
Variable speed drives and the 2/3 rule provide a mechanism to achieve these goals. In my next post, we’ll take a
look at what happens if we apply this combination of technology and technique to our example.
David Sellers
Senior Engineer – Facility Dynamics Engineering
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A Field Perspective on Engineering
Engineering lessons from the field
The “two-thirds” Rule for Locating Sensors to Control Variable Flow Systems – Part 3
Posted on November 17, 2007
In the past two posts, we have been using this simple variable flow system …
… to look at why the 2/3 rule optimizes a pumps performance as the load on the system drops.
In the first post, we looked at how the pressure requirements at different points in the system varied as the flow
varied. In the 2nd post, we looked at how allowing the load’s control valve to push the pump up its curve saved
some energy, but not nearly as much as is possible in theory. In this post, we will look at how the application of a
VFD to the pump and the application of the “2/3 rule” to the VFD control algorithm can allow us to approach the
theoretical energy savings possible as the system unloads.
As a starting point, I made a pump selection for the part load requirement of 400 gpm and 26 ft.w.c. to serve as a
baseline. (See table 1 in the 1st post to refresh your memory on where the 26 ft.w.c comes from.)
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The selection is a slightly different pump from the best efficiency selection we made for our design load of 800
gpm at 44 ft.w.c. The size of the pump is a bit smaller and its provided with an 1,150 rpm motor instead of a 1,750
rpm motor. Obviously, changing the pump every time the load on the system changes is not a practical approach.
But, this selection gives us a target to shoot for in our effort to optimize the system in a practical manner.
Here is our original pump selection, but instead of showing the performance with different size impellers, this
curve shows the performance at different speeds.
Here is the more common version of the curve with different size impellers as a frame of reference.
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If you contrast the two curves, you will notice that you can change the pumps performance by changing the
impeller or by changing the speed. But, if you change the impeller size, the efficiency of the pump tends to be
reduced as you reduce flow and/or head from the peak efficiency point or “sweet spot”.
If you change speeds, you can get the same effect in terms of a reduction in head and/or flow, but you tend to
preserve the efficiency. This is one of the advantages offered by a speed change vs. an impeller change.
Bear in mind though that if you make the speed change with a variable speed drive, the drive itself has losses.
Thus, using the drive purely as a balancing device may not be the best option when compared to an impeller trim
if only a modest change in speed is required.
Here is the variable speed curve with the system curves for a number of optimization options superimposed on it.
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Lets see what this tells us about
VFDs and the “2/3 rule”. (Note that the white line with red dots highlights our part load design flow rate target.)
The red curve is the system curve associated with allowing the control valves at the loads to push the pump up
its impeller curve with out changing speed.
Note that the flow is reduced, which saved some energy as discussed in the previous post, but the efficiency is also
reduced and the head is significantly above what is required. As a result, the control valve(s) must throttle
significantly to drive the system to the desired flow rate.
The pressure drop across the throttled control valve(s) represents an energy loss in the system (remember, flow
and head both appear in the numerator of the pump power equation). It is this loss, coupled with the reduction in
pump efficiency that causes this approach to use more energy than a pump specifically selected to deliver design
flow to one load with the load’s control valve wide open.
The orange curve shows what happens if we add a VFD to the pump and then control the VFD to maintain the
design head at the pump discharge.
As the control valve starts to throttle and push the pump up its curve, the control system senses the rise in
pressure and slows the pump down to about 1,605 rpm. As a result the control valve does not have to throttle as
hard as it would have had to if the pump did not have a VFD. This, combined with the improved pump efficiency
associated with the operating point at reduced speed makes the VFD more attractive than pushing the pump up its
curve on a pure energy savings basis.
But the head produced at the load by controlling in this manner will still be significantly above what is required
because the control point is based on the head required at the pump to deliver the design flow to the loads and we
are only delivering half of the design flow in the scenerio under discussion.
The blue curve shows what happens if we move the sensor controlling the VFD from the pump discharge to the
magical “2/3 of the way down the system” point.
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The magic is that moving the sensor away from the pump allows the sensor to “see” the pressure drop in the
piping circuit between it and the pump location. If we adjust the system to control the pressure at the 2/3 point to
the value associated with design flow at that location then, when the flow is reduced (and the head required to
move water to that location drops as the square of the reduction in flow) the system responds accordingly and
slows the pump down even further than a system with a sensor located at the pump discharge would.
The purple curveshows what would happen if we located a sensor right at the point where our load connected to
the mains and controlled to maintain the differential pressure to the load at the value required for design flow.
This location allows our optimization strategy to approach the savings that would be achieved by a pump selected
specifically for the requirements at the reduced load condition.
The table below summarizes our discussion.
Watch for the next post where I’ll
summarize our discussion with a few bottom lines before moving on to a new topic.
David Sellers
Senior Engineer – Facility Dynamics Engineering
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