ASHRAE®
Long Island
| American
Society of Heating,
Refrigerating and Air
Conditioning Engineers |
| ASHRAE Long
Island Chapter #6 |
C T T C
Case
Against Balancing Valves
Balance valves that make
it possible to adjust constant
flow systems are often used in
variable flow systems, but they
waste a great deal of energy.
This article describes this
energy loss and the calculations
for determining the energy
con-sumption of a constant speed
chilled-water system and for a
similar variable speed system
with no balance valves.
Thousands of systems
installed today prove that
variable volume, variable speed
chilled-water systems do not
require balance valves. Balance
valves in variable volume and
variable flow chilled-water
systems not only waste energy
but make these systems
difficult, if not impossible, to
control. A previous ASHRAE
Journal article1 shows that no
reason exists to install balance
valves on a variable flow
system.
Constant and Variable Flow
Systems
Figure 1 is a typical
piping sche-matic for a constant
speed, constant flow
chilled-water system
incorporating bal-ance valves
and three-way control valves on
all terminal loads such as coils
and other heat exchangers.
Figure 2 describes a typical
variable volume chilled-water
system with a bypass valve to
regulate minimum flow in the
chiller evaporator. The piping
installation shown in Figure 2
costs less than that shown in
Figure 1.
Energy
Loss and Balance Valves
The added piping cost is
minimal compared to the cost of
energy for the constant flow
system shown in Figure 1. The
total energy cost of balance
valves extends beyond the costs
associated with losses through
the valves themselves. Fig-ure 3
provides the pump flow and head
required for a variable flow
system and a similar constant
flow system. This ex-ample is
for an actual 180 ton (633 kW)
chilled-water system with a
temperature differential of 12°F
(7°C) or 2 gpm/ton (0.036 L/(s ·
kW) for a flow of 360 gpm (23
L/s). Assume that a 200 ton (703
kW) chiller was selected with a
flow of 2 gpm/ton (0.036 L/(s ·
kW) or 400 gpm (25 L/s). Figure
4 provides similar head and flow
data with a 225 ton (791 kW)
chiller selected instead. The
system then was designed for 225
ton (791 kW) at a flow of 450
gpm (28 L/s) and a total dynamic
head of 90 ft (269 kPa) when the
225 ton (791 kW) chiller was
selected.
(DSF) of around 11% for
the 200 ton (703 kW) chiller and
25% for the 225 ton (791 kW)
chiller. Evaluation of many
exist-ing systems have indicated
DSFs as high as 50%, resulting
in flows and pump heads much
higher than the actual
opera-tional flow and head. A
study of a hotel in California2
showed that a 350 ton (1231 kW)
chilled-water system was
in-stalled on a 250 ton (879 kW)
load for a DSF of 40%. Added
capacity is needed for reasons
of redundancy and future growth.
One of the great
advantages of a properly
designed, variable flow
chilled-water system is that the
system operates at the actual
load on the system, regardless
of DSF or allowances for future
expansion. In this case, as
indicated previously, the actual
design flow for the variable
flow system is 360 gpm (23 L/s)
compared to 400 gpm (25 L/s) for
the constant volume system with
a DSF of 11% and a second system
with a DSF of 25% and a flow of
450 gpm (28 L/s). The overall
DSF is a summation of all other
safety factors such as
diversity, load, and friction.
Constant
Flow and Variable Flow Systems
Operation
Most existing constant
flow systems use three-way coil
con-trol valves as shown in
Figure 1. This is the system
that is used in this evaluation
of constant flow systems.
Figures 3 and 4 show a
pump flow-head curve, a system
head curve for a variable flow
system, and a point of operation
for a constant flow system. The
design points are shown as 400
gpm (25 L/s) at 70 ft (209 kPa)
head and 450 gpm (28 L/s) at 90
ft (269 kPa), respectively. The
point of operation for all three
systems is 360 gpm (23 L/s)
because that is the actual
maxi-mum flow required for each.
The difference between the point
of operation and the design flow
is the flow reduction achieved
through the use of balance
valves on the constant flow
systems.
Compare these points of
operation for the constant flow
system to the path of operation
of a properly designed vari-able
flow, variable speed system. In
this case, the setpoint required
at the controlling differential
pressure transmitter is 10 ft
(30 kPa). This is the head
required for the branch with the
highest pressure drop.
The path of operation is
on the uniform system head curve
for the variable flow system.
The system will never reach the
design points for the constant
flow systems because no
requirement exists in the system
for these high flows and heads.
Energy
Waste of Balance Valves
The energy waste of
balance valves is due to their
use being based upon constant
flow in a system that has a
variable load. The system
operates at 360 gpm (23 L/s),
regardless of the cool-ing load
using much more energy than the
vari-able flow system shown in
Figures 3 and 4. The following
energy calculations are based
upon a chilled-water system that
operates from March to October
for an annual operation of 5,880
hours.
The same equation for
brake horsepower is used for
constant and variable speed
pumps. The basic equation is:
The data in this article
assumes that the chiller plant
is cooled by this chilled water.
If not, then the efficiency of
the mo-tors and variable speed
drives must be deducted from the
calculations.
The calculation of the
energy consumption for the
constant flow systems is simple
since the pump always operates
at one point, 360 gpm (23 L/s)
at 73.5 ft (220 kPa) head for
the system with an 11% DSF and
360 gpm (23 L/s) at 100 ft (299
kPa) head for the system with a
25% DSF. The annual energy
consumption for the constant
flow system with an 11% DSF is
40,219 kWh and 57,211 kWh for
the system with a 25% DSF. These
values were determined using
high efficiency motors with 0.8
kW (1.07 kW/kW) per brake
horsepower or 92% efficiency.
The calculations for the energy
consumption of the variable flow
system are much more complex
since they must be made at 10
points on the system head curve
along with the extra energy that
is consumed by minimum flow
through the chiller.
Tables 1 and 2 describe
these calculations with data
secured from the pump flow-head
curves of Figures 3 and 4. The
wire-to-shaft efficiencies and
percent of load in Columns E and
G are from the authors’
experience with variable primary
pumping systems. The pump energy
for maintaining minimum flow in
the chillers for the values up
to 108 gpm (7 L/s) must be added
to the total kW in Tables 1 and
2. This energy consumption is
not great since the flow and
head are at minimums. The
calculation for this energy
reveals that the energy to be
added is 264 kWh for the 11% DSF
system and 325 kWh for the 25%
DSF system.
The energy totals are
12,142 kWh for the variable flow
system with the 11% DSF and
13,528 kWh for the system with a
25% DSF. If you subtract these
totals from those for the
constant flow systems, you get
27,813 kWh saving or 68% for the
11% DSF system and 43,358 kWh or
76% for the 25% DSF system. A
summary of actual annual
measurement and verification on
a large 24/7 hotel facility in
California is available in
Reference 2. Interestingly, the
net pump energy sav-ings by
eliminating bal-ance valves and
incorporating variable speed
pumps was 80%.
The earlier figures are
for pumping energy alone and do
not include the thermal
equivalent of the pumping. An
estimate for this energy is
achieved by multiplying the
energy lost in kWh by 3,413 Btu
(3600 kJ), the heat content of a
kWh and dividing by 12,000 Btu
(12 660 kJ), the heat content of
a ton of cooling. If the overall
production of a ton of cooling
re-quires 0.9 kW/ton (0.256
kW/kW) in the chiller plant, the
added savings for the 11% DSF
system would be 7,119 kWh for a
total of 37,408 kWh and 11,099
kWh for a total of 54,457 kWh
for the 25% DSF system.
These totals are almost
unbelievable, since a
combination of energy saved in
pumping and in the chiller plant
is almost equal to the total
energy of the pumping for the
constant flow system. Since
these savings are for a 180 ton
(633 kW) system operating 5,880
hours per year, if you divide
the savings by these two
figures, you achieve a kW/ton
figure of 0.035 and 0.052 (0.01
kW/kW and 0.015 kW/kW),
respectively. This provides us
with a simple number that can be
used to estimate quickly whether
a system should be evaluated for
replacing three-way valves and
constant speed pumps with a
variable volume system.
If we round off these
numbers to 0.04, we can proceed
to make a quick estimate for a
larger system. For example, a
5,000 ton (17 585 kW) system
operating 7,000 hours/year, the
rough savings is 1.4 million kWh
per year. If the electrical rate
is $0.12/kWh, this amounts to
$168,000 per year. This is a
hypothetical case, but the
procedures used here can be used
to evaluate any existing
chilled-water system.
Assume that 20,000
chiller plants were in the U.S.
at the end of 2008 and that
these plants had a capacity of
16 mil-lion tons (5 627 200 kW).
If we also assume that half of
these plants are still constant
flow with three-way coil valves
and balance valves operating
around 5,000 hours per year,
with an overall savings factor
of 0.04 kWh per ton/hour, the
sav-ings in electric generat-ing
plant capacity would be 320 MW
and 1.6 million MWh in energy
consumption. These are just
gross estimates. However, are
they not huge enough to
encourage us to (1) stop
designing variable flow systems
with balance valves and constant
speed pumps and (2) search for
existing chilled-water systems
to find those that are
candi-dates for conversion to
variable flow systems without
balance valves?
Conclusions
Design chilled-water
systems with variable speed
pumps, high performance two-way
coil control valves, and no
balance valves. The coil control
valves must have a high shut-off
pres-sure, a broad turn-down
range, and at least a Level V
bub-ble test in accordance with
ANSI Standard 70-2-2003, Control
Valve Seat Leakage.
If you have a
chilled-water system with
three-way coil valves and
balance valves, consider
converting it to a true variable
vol-ume system. Do a quick check
by securing the annual number of
hours of operation and the
actual tons of cooling for your
system. Multiply these two
figures by 0.04 kW/ton (0.14
kW/kW) to get an estimated
annual savings in kWh for
con-verting to variable flow and
opening the balance valves wide.
Multiply your annual savings in
kWh by your composite electrical
rate to get a rough estimate of
your annual monetary savings.
This number will help you decide
whether you should do an
intensive study of your system
for conversion to variable flow.
Article In:
ASHRAE Journal, July 2009.
Please see article for all
references and credits.
By Gil
Avery, P.E., Fellow/Life
Member ASHRAE; and
James B.
(Burt) Rishel, P.E.,
Fellow/Life Member ASHRAE |
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