Energy Efficient Buildings
Variable-Air-Volume and Variable-Flow-Pumping Systems
Introduction
In the past, most heating and air conditioning systems in multi-zone buildings were designed to
deliver a constant volume of low-temperature supply air to each zone. The heating and cooling
requirements of the individual zones were met by blending cool and warm supply air or by
reheating the cool supply air. These systems are called constant-air-volume, CAV, systems.
Today, most HVAC systems are designed to vary the quantity of air to each zone. These systems
are called variable-air-volume, VAV, systems. VAV systems reduce the quantities of heating,
cooling and fan energy required to condition a building.
Similarly, in the past, most pumping systems in multi-zone buildings were designed to deliver a
constant volume of water to the end-uses independent of the actual building load. Today,
variable speed drives coupled to intelligent control vary the flow of water me match building
loads, and hence reduce pumping energy use.
This chapter shows how both constant-flow and variable-flow systems work and how to
calculate energy savings from constant to variable flow conversions.
Single-Duct CAV and VAV Systems
A single-duct commercial heating and air conditioning system is shown below. For simplicity,
the figure shows only two zones even though large commercial buildings have many zones.
Despite this simplification, building energy use can often be accurately modeled using simple
two zone models with an interior and exterior zone.
Return Air Fan
Qsen 1
Zone 1
Qsen 2
Zone 2
Qlat 1
Qlat 2
Tz1
Tz2
Reheat or VAV Box 1
TRA
Filter
Supply Air Fan
Reheat or VAV Box 2
Cooling Coil
TSA
TMA
TOA
3-Way Valve
3-Way Valve
CW Supply
CW Return
HW Supply
HW Return
3-Way Valve
HW supply
HW Return
Single-duct system with two zones.
In single-duct constant-air-volume system, SD-CAV, the flow of cool air to the zone remains
constant, but heat is added in a reheat box at each zone to meet the cooling/heating load in the
zone.
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In a single-duct variable-air-volume system, SD-VAV, zone temperature is maintained by varying
the flow of cool air down to the minimum air flow required for ventilation purposes by a VAV
box. For energy efficiency, ASHRAE recommends minimum flows of 30% of peak flow or less. If
the zone requires heating after the flow of cool air is at the minimum, then heat is added in the
VAV/reheat box. VAV boxes for an internal zone that never requires heating and an external
zone that sometime requires heating are shown in the pictures below.
VAV box without reheat
VAV box with hot-water reheat
The volume flow rate of air from the cooling coil, Vcc, and quantity of heat added in the reheat
box, Qh, for as functions of zone temperature are shown below.
Vcc
Qh
Qh
Vcc
Vmin
Tsp
Tzone
SD-CAV
Deadband
Tzone
SD-VAV
SD-CAV Heating and Cooling Energy Use
In a SD-CAV system, the flow of supply air through the cooling coil, Vsa, is constant, and the
enthalpy, hsa, and temperature, Tsa, of supply air leaving the cooling coil are known. Thus, the
energy extracted by the cooling coil, Qcc, can be calculated from an energy balance on the
cooling coil, based on the enthalpy, h, and density, p, of the mixed air, ma, and supply air, sa and
water, w.
Qcc = Vsa psa (hma – hsa) - Vsa psa (wma – wsa) hw
2
The heating energy added by the heating coil in reheat box 1, Qhc1, can be calculated from an
energy balance on the zone and reheat box, based on product of the air density and specific
heat pcpa.
Qhc1 = Vsa1 pcpa (Tra – Tsa) – Qsen1
The total building heating energy requirement is the sum of the heating energy required by each
reheat box.
SD-VAV Heating and Cooling Energy Use
In a SD-VAV system, the flow of supply air through the cooling coil, Vsa, varies and is the sum of
the air flow required by each zone. The procedure is:
Vsa1 = Qsen1 / (pcpa (Tra – Tsa))
If Vsa1 < Vsa1,min then ‘heating required
Vsa1 = Vsa1,min
Qhc1 = Vsa1 pcpa (Tra – Tsa) – Qsen1
Else ‘no heating required
Vsa1 = Qsen1 / (pcpa (Tra – Tsa))
Qhc1 = 0
End If
The total building heating energy requirement is the sum of the heating energy required by each
VAV/reheat box. The total air flow over the cooling coil, Vsa, is the sum of the air flows through
each zone. And, as before, the energy extracted by the cooling coil, Qcc, can be calculated from
an energy balance on the cooling coil.
Qcc = Vsa psa (hma – hsa) - Vsa psa (wma – wsa) hw
Dual-Duct CAV and VAV Systems
A dual-duct commercial heating and air conditioning system is shown below. For simplicity, the
figure shows only two zones even though large commercial buildings have many zones. Despite
this simplification, building energy use can often be accurately modeled using simple two zone
models with an interior and exterior zone.
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Return Air Fan
Qsen 1
Zone 1
Qlat 1
CW Return
CW Supply
RA
Qsen 2
Zone 2
Qlat 2
Tz1
Tz2
3-Way Valve
Mixing or VAV Box 1
Mixing or VAV Box 2
CSA
Cooling Coil
MA
OA
Heating Coil
HSA
Filter
Supply Air Fan
3-Way Valve
HW Supply
HW Return
Dual-duct system with two zones.
In a dual-duct constant-air-volume system, DD-CAV, warm and cool air streams are mixed in a
mixing box at each zone to meet the cooling/heating load in the zone. Note that the total air
flow to the zone remains constant. In a dual-duct variable-air-volume system, DD-VAV, zone
temperature is maintained by varying the amount of cold or warm air introduced into the zone.
When the zone calls for heating, all cooling air is shut off. When the zone calls for cooling, all
heating air is shut off. The warm and cold air streams are mixed only during very low-load
conditions, to maintain a minimum air flow into the zone. The result is that the total flow of air
into the zone varies, rather than remaining constant as in CAV systems. The volume flow rate of
air from the cooling coil, Vcc, and heating coil, Vhc, as functions of zone temperature are shown
below.
Vcv
Vvav
Vhc
Vcc
Vhc
Vcc
Vmin
Tsp
DD-CAV
Tzone
Tsp
Tzone
DD-VAV
DD-CAV Heating and Cooling Energy Use
In a DD-CAV system, the flow of supply air through the supply air fan, Vsa, is constant; however,
the flow rate through the cooling and heating coils, Vcsa and Vhsa, varies. The fraction of supply
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air to each zone from the cooling coil, fcc, can be determined from an energy balance on each
zone and mixing box.
fcc1 = [Qsen1 / (Vsa1 pcpa) + (Thsa- Tra)] / [Thsa – Tcsa]
The flow of air through the cooling and heating coils for each zone, Vcsa and Vhsa, can then be
calculated.
Vcsa1 = fcc1 x Vsa
Vhsa1 = (1-fcc1) x Vsa
The total air flows through the cooling and heating coils, Vcsa and Vhsa, are the sum of the flows
from each zone. The cooling and heating coil energy use, Qcc and Qhc, can then be calculated
from energy balances on each coil.
Qcc = Vcsa psa (hma – hsa) - Vsa psa (wma – wsa) hw
Qhc = Vhsa pcpa (Tsa – Tma)
DD-VAV Heating and Cooling Energy Use
In a DD-VAV system, the flow of supply air through the cooling coil, Vsa, varies and is the sum of
the air flow required by each zone. The procedure to calculate savings begins by calculating the
cooling and heating air flows through each zone from an energy balance on the zone and mixing
box.
If Qsen1 > 0 then ‘cooling
Vhsa1 = 0
Vcsa1 = [-Qsen + Vha1 pcp Tra – Vhsa1 pcpa Thsa] / [pcpa Tcsa]
If Vcsa1 < Vsa,min1 then ‘use method for DD-CAV, but with Vsa1 = Vsa,min1
fcc1 = [Qsen1 / (Vsa,min1 pcpa) + (Thsa- Tra)] / [Thsa – Tcsa]
Vcsa1 = fcc1 x Vsa,min1
Vhsa1 = (1-fcc1) x Vsa,min1
End if
Else ‘heating
Vcsa1 = 0
Vhsa1 = [-Qsen + Vsa1 pcp Tra – Vcsa1 pcpa Tsa] / [pcpa Thsa]
If Vhsa1 < Vsa,min1 then ‘use method for DD-CAV, but with Vsa1 = Vsa,min1
fcc1 = [Qsen1 / (Vsa,min1 pcpa) + (Thsa- Tra)] / [Thsa – Tcsa]
Vcsa1 = fcc1 x Vsa,min1
Vhsa1 = (1-fcc1) x Vsa,min1
End if
End if
The total air flows through the cooling and heating coils, Vcsa and Vhsa, are the sum of the flows
from each zone. The cooling and heating coil energy use, Qcc and Qhc, can then be calculated
from energy balances on each coil.
Qcc = Vcsa psa (hma – hsa) - Vsa psa (wma – wsa) hw
Qhc = Vhsa pcpa (Tsa – Tma
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Fan and Pump Curves
The volume flow rate generated by a fan depends on the total system pressure drop. Fans can
generate high volume flow rates at low system pressure drops or low volume flow rates at high
system pressure drops. A fan curve shows the relationship between total pressure drop and
volume flow rate for a specific fan.
Static Pressure
Static Efficiency
Modified graph from ASHRAE Handbook: HVAC Systems and Equipment 2008
It is common for fan manufacturers to publish fan performance data in terms of “static
pressure” versus flow. The “static pressure’ in performance data is actually the difference
between the static pressure at the fan outlet and the total pressure at the fan inlet.
Pstatic,performance data = Pstatic,2 – (Pstatic,1 + Pvelocity,1)
All fan performance calculations should be performed using total pressure. For example, the
methods to calculate pressure loss through ducts and fittings, calculate total pressure loss. In
addition, the power requirement of a fan is a function of total pressure loss, not static pressure
loss. Thus, for fan calculations, it is important to add the velocity pressure of the air leaving the
fan outlet to the static pressure reported in fan performance data to determine the relationship
between total pressure and flow for the fan. In some cases manufactures list outlet velocity for
the fan. In other cases, outlet velocity can be calculated from outlet dimensions of the fan and
airflow. The total pressure of the fan is then:
Ptotal = Pstatic,performance data + Pvelocity,2 = (Pstatic,2 + Pvelocity,2 )– (Pstatic,1 + Pvelocity,1)
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In U.S. dimensional units, the relation is:
Ptotal (in-H20) = Pstatic,performance data (in-H20) + (V2 / 4,005)2 where V2 is in ft/min
System Curve
The total pressure rise that the fan must produce to move air is determined by the duct system.
This total pressure of the duct system is the sum of the pressure due to inlet and outlet
conditions and the pressure loss due to friction. In a duct system, pressure loss due to friction
increases with increasing fluid flow; thus, system curves have positive slopes on pump
performance charts. The operating point of a fan is determined by the intersection of the fan
and system curves.
The equations for pressure loss from friction through ducts and through fittings are:
Pp = (f L fluid V2) / (2 D)
Pf = kf fluid V2 / 2
These equations clearly show that for a given duct system, the pressure drop is proportional to
the square of the velocity, and hence the square of the volume flow rate.
Pfriction = C1 V2 = C2 V2
This quadratic relationship can be plotted on the fan curve to show the “system curve”. The
operating point of the fan will be at the intersection of the fan and system curves, as shown
below.
Modified graph from ASHRAE Handbook: Systems 2008
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Plotting System Curves
A system curve for a duct system with negligible inlet/outlet pressure differences is a parabola
of the form hheadloss = C2 V2. The curve passes through the origin because the inlet/outlet
pressure difference, sometimes called the static head, is zero. The coefficient C2 can be
determined if the operating point is known by substituting the known pressure drop and flow
rate into the equation and solving for C2. The fluid work required to push the fluid through the
duct is the product of the volume flow rate and system pressure drop and is represented
graphically by the area under the rectangle defined by the operating point.
Controlling Flow Using Throttling/Outlet Dampers:
Controlling flow by closing a flow-control valve/damper downstream of the pump/fan increases
pressure drop and causes the operating point to move up and left on the pump/fan curve.
Pump Curve
Head
Throttled
System Curve
Flow Rate
This results in relatively small energy savings, since
Wf2 = V2 P2
where
V2 < V1 but
P2 > P1
Thus, throttling is an energy inefficient method of flow control.
Controlling Flow by Reducing Fan Speed
The most energy efficient method of varying flow is by controlling pump/fan speed. If the
required flow varies over time, speed control is best facilitated by an electronic variable speed
drive (VSD) which can continuously and smoothly adjust pump/fan speed as needed. One time
reductions in flow are more cost effectively accommodated by replacing the pump/fan pulley
with a larger diameter pulley to slow pump/fan rotation or by reducing pump impellor diameter.
The following example illustrates energy savings from reducing flow by slowing fan speed.
The figure below shows fan performance at various speeds and a system curve. At full speed
with no throttling, the fan operates at B. At B, the pressure is 6.25 in-H20, the volume flow rate
is 7,600 cfm and the required power to the fan would reduce to about 12.8 hp. If the flow is
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throttled, the fan would operate at point A. The volume flow rate at A would be about 6,700
cfm and the system pressure drop would be about 7.5 in-H20. From the chart, the required
power to the fan would be about 13 hp.
To realize significant energy savings at low flow, it is necessary to slow the fan speed. To
determine the fan speed required to deliver the initial air flow of 6,700 cfm, it is necessary to
develop a system curve for the new duct system. Pressure drop always varies with the square of
flow rate. Thus, the equation for a system curve can be written as:
h = C V2
The coefficient, C, for the new system curve can be found by substituting the values of pressure
drop and volume flow rate for point B.
C = h / V2 = 6.25 / 76002 = 1.082 x 10-7
Thus the pressure drop through the new duct system at 6,700 cfm would be about:
h = C V2 = 1.082 x 10-7 (6700)2 = 4.86 in-H20
According to the fan curves, the fan would deliver 6,700 cfm at 4.86 in-H20 if the fan speed were
slowed to about 2,500 rpm at point C. At this operating point, the fan would require about 9 hp
of power. Thus, the savings from reducing flow by slowing the fan would be about:
9
13 hp – 9 hp = 4 hp
Controlling Flow by Reducing Pump Speed
The most energy efficient method of varying flow is by controlling pump/fan speed. If the
required flow varies over time, speed control is best facilitated by an electronic variable speed
drive (VSD) which can continuously and smoothly adjust pump/fan speed as needed. One time
reductions in flow are more cost effectively accommodated by replacing the pump/fan pulley
with a larger diameter pulley to slow pump/fan rotation or by reducing pump impellor diameter.
The following example illustrates energy savings from reducing flow by slowing pump/fan
speed.
The figure below shows pump performance at various speeds and a system curve.
Assume the original operating point, A, is 1,200 gpm at 55 ft-H20. According to the chart, the
required power to the pump at this operating point is about 23 hp. Alternately, pump power
could be calculated as:
WA = 1,200 gpm x 55 ft-H20 / (3,960 gpm-ft-H20/hp x 0.74) = 22.6 hp
If the flow were reduced to 900 gpm with a flow control valve, the operating point would move
along the pump curve to 900 gpm at 62 ft-H20. According to the chart, the required power to
the pump at this operating point is about 20 hp. Alternately, pump power could be calculated
as:
W = 900 gpm x 62 ft-H20 / (3,960 gpm-ft-H20/hp x 0.70) = 20.1 hp
10
Thus, the pump power savings from reducing flow from 1,200 gpm to 900 gpm with a flowcontrol valve would be about:
22.6 hp – 20.1 hp = 2.5 hp
Alternately, the flow could be reduced from 1,200 gpm to 900 gpm by slowing the pump speed
with a VSD. Reducing the pump speed from 1,200 rpm at point A to 900 rpm at point B would
reduce the volume flow rate from 1,200 gpm to 900 gpm. The reduced volume flow rate would
also generate less friction, and the system pressure drop would be reduced from 55 ft-H20 to 30
ft-H20. The power required to pump a fluid is the product of the volume flow rate and pressure
drop; hence, the areas enclosed by the rectangles defined by each operating point represent the
fluid power requirements, WA and WB, at the different flow rates.
WA = 1,200 gpm x 55 ft-H20 / 3,960 gpm-ft-H20/hp = 16.7 hp
WB = 900 gpm x 30 ft-H20 / 3,960 gpm-ft-H20/hp = 6.8 hp
The power, WB, required by the pump at point B can be read from the chart to be about 10 hp.
Alternately, the power could be calculated as:
WB = 900 gpm x 30 ft-H20 / (3,960 gpm-ft-H20/hp x 0.67) = 10.1 hp
Pump power savings would be the difference between PA and PB.
Savings = PA – PB = 22.6 hp – 10.1 hp = 12.5 hp
Alternately, power savings from reducing the volume flow rate can be estimated from the pump
affinity laws. Theoretically, pump work varies with the cube of volume flow rate. Use of the
cubic relationship would predict:
PB = PA (VB/VA)3 = 22.6 hp x (900 gpm / 1200 gpm) 3 = 9.5 hp
The 9.5 hp predicted by the pump-affinity law is less than the 10.1 hp predicted by the pump
curve. This example demonstrates how use of the cubic relationship typically exaggerates
savings. In practice, the efficiencies of the VSD, pump and motor typically decline as flow rate
decreases, resulting in slightly less savings than would be predicted using this ‘cubic’
relationship. Thus, we estimate that pump/fan work varies with the 2.5 power of flow rather
than the cube of flow. Using this relationship, if we measured PA to be 22.6 hp at 1,200 gpm, we
would estimate PB for 900 gpm to be about:
PB = PA (VB/VA)2.5 = 22.6 hp x (900 gpm / 1200 gpm) 2.5 = 11.0 hp
Thus, savings would be about:
Savings = PA – PB = 22.6 hp – 11.0 hp = 11.6 hp
This slightly lower estimate of savings incorporates the reduction in motor efficiency, and power
loss by the VSD.
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Fan/Pump Affinity Laws
The fundamental fluid mechanic relationships developed thus far can be modified to generate
other useful relations between fan parameters. These relationships are known as fan affinity
laws. The two most important relationships are derived below.
As shown in the section of system curves, friction head loss is proportional to the square of the
volume flow rate.
Pfriction = C1 V2 = C2 V2
By substitution, fluid work is proportional to the cube of volume flow rate
Wf = V Pfriction = V C2 V2 = C2 V3
Since Wf / V3 is constant, it follows that:
(Wf / V3)1 = C = (Wf / V3)2
Wf2 = Wf1 (V2 / V1)3
This relation shows that a small reduction in the volume flow rate results in a large reduction in
the fluid work. For example, reducing the volume flow rate by one half reduces fluid work by
88%!
Wf2 = Wf1 (1/2)3 == Wf1 (1/8)
(Wf1 – Wf2) / Wf1 = [Wf1 - Wf1 (1 /8)] / Wf1 = 1 – (1/8) = 88%
Another useful relation can be derived from the relationship between volume flow rate V and
the rotational speed of the pump fan. In centrifugal pumps and fans, the volume flow rate is
proportional to the rotational speed of the pump fan.
V = C RPM
Since V/RPM is constant, it follows that:
(V / RPM)1 = C = (V / RPM)2
V2 = V1 (RPM2 / RPM1)
Thus, volume flow rate varies in proportion to pump/fan speed.
VAV Control
A VFD installed on the AHU fan motor typically uses a pressure sensor signal to control the
speed of the fan. The principle types of VFD control are:
Fan-Outlet Control: In fan-outlet control, the pressure sensor is located at the outlet of the
fan and sends a signal to the VFD to maintain a duct static pressure set-point at design
conditions.
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Supply-Duct Control: In supply-duct control, the pressure sensor is typically downstream in
the duct system. The pressure sensor sends a signal to the VFD to maintain a set static
pressure at this location to meet the design pressure requirement of the terminals
downstream of the sensor.
Critical-Zone Control: In critical-zone control, the sensor can be located anywhere in the
supply air duct (but typically at the fan outlet); the duct static pressure set-point is
dynamically reset to meet the flow requirement of the most critical zone at any moment.
The location of the pressure sensors is depicted in the simple duct system shown below with
three zones (1, 2 and 3) served by VAV boxes.
B
Damper
Coil
1
C
2
3
A
In fan-outlet control, the pressure sensor controlling the VFD is located in position A. In supplyduct control, the pressure sensor controlling the VFD is located at B or C. Location B represents
the common rule of thumb of 2/3 of the distance down the duct. Location C represents the
location immediately upstream of the most distant zone.
Fan and system curves for different types of control are plotted in Figure 1. When the fan is
running at full speed, a design system flow rate of Q0 can be met at a system pressure of P0.
Fluid power is the area of the rectangle defined by the pressure drop and flow rate. To visualize
fan power requirements as a function of control types and flow rate, consider a change in
system flow rate from Q0 to Q’. In an AHU without a VFD the closing of dampers at VAV
terminals shifts the system curve, which raises the pressure from P0 to P1 on the fan curve and
reduces system flow from Q0 to Q’. The fan power requirement is reduced, but not
proportionally with the reduction in flow rate since pressure drop increases.
In fan-outlet control with a VFD, the duct static pressure is maintained at P0. Fan power
requirement is reduced proportionally with the reduction in flow rate because the VFD slows
the fan speed.
Savings can be increased by controlling VFD fan speed to maintain static pressure near the most
distant VAV box(es). This is called supply-duct control. However, supply-duct control risks
starving VAV boxes in parallel branches whose pressure requirement exceeds the requirement
of the branch where the pressure sensor is located. To mitigate this risk, the pressure sensor is
often placed upstream of parallel branches. A common rule of thumb is to place the sensor 2/3
of the way down the longest duct; however, this location negates energy savings from zones
downstream of the sensor and results in an artificially high static pressure.
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Critical zone control eliminates this risk by identifying the zone with the greatest duct static
pressure requirement at its VAV terminal. In critical-zone control, the duct static pressure setpoint is constantly adjusted to meet the flow requirement of the most critical zone; therefore,
no minimum pressure needs to be maintained and the system curve can approach zero system
pressure when system flow rate approaches zero. Thus, critical-zone control eliminates the risk
of starving VAV boxes, and results in greater fan power reduction than supply duct control when
the pressure sensor is placed upstream of parallel zones.
In summary, fan-outlet control allows the fan to slow down to a new fan curve while
maintaining the set-point pressure of P0. Supply-duct control allows the system pressure to be
reduced to P2. Critical-zone control allows the system pressure to be greatly reduced to P3. Since
the system flow rate is determined by the terminal unit controllers through the actions of
terminal dampers and the AHU system efficiency is not directly controllable, minimizing the duct
static pressure set-point is the primary strategy to minimize AHU power. Among the three
common control methods, critical-zone control results in the lowest system pressure at any flow
requirement and thus offers the greatest AHU power savings.
Figure 1. System and fan curves for AHUs without VFD, with fan-outlet control, with supply-duct
control and with critical-zone control illustrate the total system power at reduced system flow.
The implementation of critical zone reset control depends on the particular constraints of a
system. Critical-zone control methods depend on collecting information from the VAV terminal
units and using the information in an algorithm to dynamically reset the static pressure setpoint. Zheng and Liu (2005) proposed resetting the chiller pump VFD set-point pressure based
on the pressure drop across the identifiable most resistant loop (MRC). In most commercial
buildings, however, the most critical zone changes constantly depending on thermal loads in
each zone and thus must be algorithmically identified in the control strategy. Hartman (1993), in
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his Terminal Regulated Air Volume (TRAV) algorithm, described a method to vary fan speed to
meet flow requirement in each zone. Warren and Norford (1993) improved upon the TRAV
algorithm by resetting the static pressure set-point to minimize the number of starved VAV
terminal units. Moult (1999) proposed an algorithm to modulate fan speed in order to maintain
the damper position of the most open VAV terminal in a specified range.
Ma et al. (2014) proposed a new algorithm that reduced to modulate fan speed. The proposed
strategy uses the number of open terminal dampers as a control input, allowing larger
adjustments to the static pressure set-point and maintaining the functionality of the controls
even when a terminal damper malfunctions. The control strategy is computationally simple and
utilizes data readily available on most current and many older BAS.
The control loop sequence is illustrated in Figure 2. The algorithm executes the following steps
in a loop:
1. Poll through all terminal unit controllers and determine the number of terminals with
damper position greater than Pos_open.
2. If the number of open terminals is greater than N_max and the duct static pressure is
less than P_max, increase the duct static pressure set-point by delP.
3. Else if the number of open terminals is less than N_min and the duct static pressure is
greater than P_min, decrease the duct static pressure set-point by delP.
4. Else, maintain current duct static pressure ste-point.
5. Delay by an amount of time defined in t_delay.
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Figure 2. Algorithm to execute the proposed control strategy
Fan Energy Use and Control Efficiency in VAV Systems
The power required to move air through a duct system is:
Power = P x V / 
Where, ΔP is the pressure rise across the fan, V is the system flow rate, and η is the total
efficiency of the fan, drive, motor and VFD. As shown in the preceding sections, in VAV systems,
each VAV box modulates air flow to the minimum needed by the zone it serves. Thus, the total
system flow rate V is sum of the required flow rates to each VAV box. Because VAV systems
require substantially lower volume air flow rates than CAV systems, fan energy savings can be
significant.
However, to minimize power, a control system must also minimize the pressure rise across the
fan P to the lowest achievable to meet the required flow rate V. The most energy efficient
method of reducing P is by slowing fan speed with a variable frequency drive (VFD). In an ideal
system, the pressure will be decreased along the design system curve as the flow rate
decreases. When this occurs, the fluid work supplied by a fan or pump varies with the cube of
volume flow rate. In this case, the fluid work supplied by the fan, Wf2, at reduced volume flow
rate Vsa2 is the product of the original fluid work supplied by the fan, Wf1, at the original
volume flow rate, Vsa1, and the cube of the ratio of the reduced and original flow rates:
Wf2 = Wf1 (Vsa2 / Vsa1)3
Unfortunately, this ideal condition is difficult to achieve in real VAV systems in which a single fan
serves multiple zones. Depending on the type of control and location of required flow rates to
each zone, the actual fluid work produced by the fan is often much greater than this ideal
minimum. The ratio of actual to minimum fan power due to the type of control and location of
required flow rates is called the VAV control efficiency.
Control Efficiency and Air Flow in Serial Duct Systems
To understand VAV control efficiency, consider the simple duct system shown below with three
zones (1, 2 and 3) served by VAV boxes.
B
Damper
Coil
1
C
2
3
A
In fan-outlet control, the pressure sensor controlling the VFD is located in position A. In supplyduct control, the pressure sensor controlling the VFD is located at B or C. Location B represents
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the common rule of thumb of 2/3 of the distance down the duct. Location C represents the
location immediately upstream of the most distant zone. The table below shows control
efficiencies, e_cntrl, for pressure sensor locations A, B and C in a single duct system. Control
efficiencies are calculated for the design flow rate of 7,500 cfm and for different combinations
of VAV box flow rates that result in a total flow of 5,000 cfm. The table shows that, regardless of
the flow scenario, control efficiency increases as the sensor is located farther downstream. The
table also shows that even when the pressure sensor is located at the most distant position C,
control efficiency is less than 100% whenever the flow through the last zone (3) is less than the
design (maximum) flow rate.
V1 (cfm)
V2 (cfm)
V3 (cfm)
Pressure Setpoint at
Control Location A
Design
Case
2,500
2,500
2,500
V2 & V3 V1 & V3 V1 & V2 Balanced
Reduced Reduced Reduced Reduction
2,500
1,250
1,250
1,667
1,250
2,500
1,250
1,667
1,250
1,250
2,500
1,667
4.00
4.00
4.00
4.00
4.00
e,cntrl = dh,act / dh,max
Pressure Setpoint at
Control Location B
1.00
0.64
0.69
0.82
0.67
3.00
3.00
3.00
3.00
3.00
e,cntrl = dh,act / dh,max
Pressure Setpoint at
Control Location C
1.00
0.72
0.78
0.92
0.75
2.00
2.00
2.00
2.00
2.00
e,cntrl = dh,act / dh,max
1.00
0.82
0.85
1.00
0.83
In critical-zone control, the sensor can be located anywhere in the supply air duct, but is
typically located at the fan outlet. The duct static pressure set-point is dynamically reset to
meet the flow requirement of the most critical zone at any moment. Thus, by design, criticalzone control achieves 100% control efficiency and is the most efficient type of VAV control.
Control Efficiency and Air Flow in Parallel Duct Systems
The advantages of critical-zone control become even greater in more complicated duct systems
with parallel flow. For example, consider the duct system shown below with two parallel ducts
and six zones.
17
B
Damper
C
AB
Coil
1
2
3
A
AD
D
4
E
5
6
If the pressure sensor were located at C, then it is possible to starve zones 4, 5 and 6 by not
providing sufficient pressure at the fan for these zones. Starving can occur when zones 1, 2 and
3 require minimum air flows and zones 4, 5 and 6 require high air flows; the sensor at C can’t
detect the high flow, and hence, pressure requirement in the other duct. Unfortunately, this
type of unbalanced flow between parallel supply ducts is common; for example, it can be caused
by the sun moving from one exposure of the building to the other exposure.
The table below shows control efficiencies and required flow rates and actual flow rates for
different flow scenarios in the parallel duct system. In this simulation, the total air flow
requirement is reduced from the design condition of 15,000 cfm to 12,500 cfm. As before, the
table demonstrates that control efficiency increases as the control location moves downstream.
But it also shows that downstream control locations increase the probability of starving zones in
the parallel ducts. Flow rates to starved zones are highlighted and their corresponding control
efficiencies labelled as not applicable (N.A.). Reducing the probability of starved zones is one
reason why VAV control designers often move the pressure sensor upstream of clusters of
parallel ducts, albeit at the expense of reduced control efficiency. The 100% control efficiency
and minimal probability of starving zones make critical-zone control highly attractive.
18
Design
V1 (cfm)
V2 (cfm)
V3 (cfm)
V4 (cfm)
V5 (cfm)
V6 (cfm)
Pressure Setpoint at
Control Location A
e,cntrl = dh,act / dh,max
V4_actual (cfm)
V5_actual (cfm)
V6_actual (cfm)
Pressure Setpoint at
Control Location B
e,cntrl = dh,act / dh,max
V4_actual (cfm)
V5_actual (cfm)
V6_actual (cfm)
Pressure Setpoint at
Control Location C
e,cntrl = dh,act / dh,max
V4_actual (cfm)
V5_actual (cfm)
V6_actual (cfm)
Case
2,500
2,500
2,500
2,500
2,500
2,500
V2 & V3 V1 & V3 V1 & V2 V1 & V4
V3 & V6
Balanced
Balanced
Reduction in Reduction in
Reduced Reduced Reduced Reduced Reduced Zone 1,2,3 Zone 4,5,6
2,500
1,250
1,250
1,250
2,500
1,667
2,500
1,250
2,500
1,250
2,500
2,500
1,667
2,500
1,250
1,250
2,500
2,500
1,250
1,667
2,500
2,500
2,500
2,500
1,250
2,500
2,500
1,667
2,500
2,500
2,500
2,500
2,500
2,500
1,667
2,500
2,500
2,500
2,500
1,250
2,500
1,667
9.25
9.25
9.25
9.25
9.25
9.25
9.25
9.25
1.00
2,500
2,500
2,500
0.77
2,500
2,500
2,500
0.77
2,500
2,500
2,500
0.77
2,500
2,500
2,500
0.68
1,250
2,500
2,500
0.60
2,500
2,500
1,250
0.77
2,500
2,500
2,500
0.77
1,667
1,667
1,667
3.00
3.00
3.00
3.00
3.00
3.00
3.00
3.00
1.00
2,500
2,500
2,500
N.A.
2,500
2,500
1,449
N.A.
2,500
2,500
1,449
N.A.
2,500
2,445
1,543
1.00
1,250
2,500
2,500
0.89
2,500
2,500
1,250
N.A.
2,500
2,357
1,687
1.00
1,667
1,667
1,667
2.00
2.00
2.00
2.00
2.00
2.00
2.00
2.00
1.00
2,500
2,500
2,500
N.A.
2,500
2,250
1,069
N.A.
2,500
2,440
1,076
N.A.
2,500
2,412
1,133
1.00
1,250
2,500
2,500
0.95
2,500
2,500
1,250
N.A.
2,500
2,020
1,665
1.00
1,667
1,667
1,667
CAV to VAV: Energy Saving Recommendation Example
The methods for calculating savings shown above can be easily incorporated into computer
program, such as ESim, which calculates energy use on an hourly data and sums the results for
an entire year. The case study example shown here demonstrates the use of ESim to estimate
savings from switching from a CAV to VAV system.
Recommendation
We recommend that retrofitting the air distribution system from a DD-CAV to a DD-VAV system.
This would entail replacing the reheat boxes at each zone with VAV boxes with damper controls
and adding VSDs to each supply air fan.
Estimated Savings
Building HVAC energy use is affected by outside air temperature, outside air humidity, solar
radiation, heat transfer through the building envelope, internal electricity use, heat gain from
building occupants, the efficiency of chillers and boilers, and the type and control of the air
distribution system. Building energy simulation software attempts to model these complex
interactions and estimate building energy use.
We modeled energy use in the administration and engineering buildings using the ESim building
energy simulation software (Kissock, 1997). ESim uses a description of the building and its
19
energy using equipment with hourly meteorological data from TMY2 data files (NREL, 1990) to
predict hour-by-hour building energy use. Figure 3.2 below shows simulated and actual
electricity use for the engineering and administration buildings. Figure 3.3 shows simulated and
actual steam use for the engineering and administration buildings. The close match between
simulated and measured energy consumption indicates that the model relatively accurately
captures the major energy interactions within the buildings.
Figure 3.2 Simulated and actual electricity use for the engineering and administration buildings
with CAV air distribution systems.
20
Figure 3.3 Simulated and actual steam use for the engineering and administration buildings with
CAV air distribution systems.
Next, we simulated building energy use with a VAV system instead of a CAV air distribution
system. The predicted energy consumption with each system, and the savings from switching to
a VAV system, are shown in Table 3.1. The large reduction in steam energy use suggests that in
most zones internal electricity use and heat from the building occupants is sufficient to heat the
zone most of the time. The relatively small reduction in electrical energy use suggests that the
current practices of slowing down the fans and employing an economizer cycle successfully
reduce fan energy use and mixing of hot and cold air streams; thus, electricity savings are
relatively small.
Table 3.1 Simulated energy use and savings from CAV and VAV systems.
Electricity
Steam
(kWh/yr)
(1000 lb/yr)
DD-CAV
9,297,133
11,277
DD-VAV
8,899,816
2,255
Savings
397,317
9,022
Percent Savings
4.3%
80%
To estimate cost savings, we assume that steam costs $6 per thousand pounds and use the
average cost of electricity of 4.8 cents per kWh. Based these simulation results, estimated
savings from converting to a DD-VAV systems would be about:
9,661,149 kWh/yr x 4.3% = 415,429 kWh/yr
415,429 kWh/yr x $0.048 /kWh = $19,941 /yr
21
12,547 ThousLb/yr x 80% = 10,038 ThousLb/yr
10,038 ThousLb/yr x $6 /ThousLb = $60,266 /yr
$19,941 /yr + $60,266 /yr = $80,167 /yr
Estimated Implementation Cost
Management and maintenance personnel roughly estimate that converting the CAV zone mixing
boxes to VAV zone mixing boxes would cost about $500 per mixing box. Roughly assuming that
there are 100 mixing boxes in the building, the conversion cost would be about:
100 mixing boxes x $500 /mixing box = $50,000
VSD suppliers estimate that installed cost of a VSD for a 100-hp supply fan is about $10,000, and
the installed cost of a VSD for a ~50-hp supply fan is about $6,000. Thus, the cost of VSDs for
the supply fans would be about:
(6 x $10,000) + (3 x $6,000) = $78,000
Depending on the design, it may be possible to use the current return-air fans without VSDs.
Based on these estimates, the total implementation cost would be about:
$50,000 + $78,000 = $128,000
Estimated Simple Payback
$128,000 / $80,167 /yr x 12 mo/yr = 19 months
22
CAV to VAV: Measured Savings Example 1
The graphs below show electricity use, chilled water energy use and hot water energy use
before and after a CAV to VAV retrofit of the Texas A&M University Zachry Engineering Center.
23
CAV to VAV: Measured Savings Example 2
The graphs below show electricity use, chilled water energy use and hot water energy use
before and after a CAV to VAV retrofit of a building at the University of Texas.
24
VSD Pumping: Commercial Building Example
A typical constant-flow commercial-building chilled water system is shown below. This system
provides constant flows through the condenser and evaporator of each chiller whenever a
chiller is operational. The primary chilled water pumps are typically much smaller than the
secondary pumps, since a primary pumps only have to move water through the chiller
evaporators and not through the entire building. The secondary chilled water pumps are much
larger than the primary pumps and provide a constant flow of chilled water to the AHUs. Each
AHU varies the quantity of chilled water through the coil and bypasses unneeded chilled water.
Cooling Tower Fan
Chilled Water Supply
AHU 1
Chiller 1
Chiller 2
Condenser
(Cooling Tower)
Pumps
AHU 2
AHU 3
Secondary
Chilled Water
Pumps
Chilled Water Return
Primary
Chilled Water
Pumps
Close-ups of a typical piping configuration at the air handler cooling coils in a constant-flow
chilled-water supply system are shown below. The three-way valves direct chilled water either
through the cooling coil or around the cooling coil via the bypass loop. The flow of chilled water
through the cooling coils is varied to maintain the temperature of the air leaving the cooling
coils at a constant temperature. In a VSD retrofit, the bypass valves would be closed, and a
differential-pressure sensor would be installed between the supply and return headers at the air
handler located farthest from the pump. In some cases, it may be necessary to replace the
three-way valves with two-way valves if the three-way valves were not designed to handle
larger pressure drops in a VSD situation.
25
Cooling Coil
Manual
2-Way Valve
Tsa
Automatic
3-Way Valve
Cooling Coil
Manual
2-Way Valve
Tsa
Automatic
3-Way Valve
Chilled Water Supply
Chilled Water Return
Piping configuration at air handling units.
The greatest pump energy savings come from changing the secondary chilled-water loop from
constant to variable flow. This is done by:




Removing or blocking the bypass piping on each AHU
Replacing 3-way valves with 2-way valves on each AHU
Adding VFDs to the secondary chilled water pumps
Controlling the VFDs based on the differential pressure between the supply and return
headers
A typical variable flow secondary chilled water loop system is shown below.
Cooling Tower Fan
VFD
Chilled Water Supply
VFD
AHU 1
AHU 2
AHU 3
Chiller 1
dP
Secondary
Chilled Water
Pumps
Chiller 2
Condenser
(Cooling Tower)
Pumps
Chilled Water Return
Primary
Chilled Water
Pumps
26
Modern chillers are designed to accommodate variable flow through the evaporators and
condensers. This enables full variable flow chilled water plants. A variable-flow chilled water
plant with a flow control and bypass valve to guarantee minimum flow to the chillers is shown
below.
Cooling Tower Fan
VFD
Chilled Water Supply
AHU 1
AHU 2
AHU 3
Chiller 1
dP
VFD
VFD
Chiller 2
VFD
Condenser
(Cooling Tower)
Pumps
Bypass
Valve
Flow
Meter
Chilled Water Return
VFD
Primary
Chilled Water
Pumps
The savings from constant to variable flow are represented in the Figure below.
In practice, actual savings depend on DP sensor location
27
Bypass Valve
ΔP
ΔP
Process Lines
ΔP
ΔP
VSD
Chiller
Pump
Savings also depend on the quantity of bypass flow.
Bypass Valve
ΔP
Process Lines
VSD
Chiller
Pump
Savings also depend on the dP setpoint. Setting it too high, reduces savings.
28
Bypass Valve
ΔP
Process Lines
VSD
Chiller
Pump
Case Study
References
ASHRAE Handbook: Fundamentals, 1977 and 1985, ASHRAE.
Engineering Design Reference Manual, 1990, United McGill.
Kreider and Rabl, 1994, Heating and Cooling of Buildings, McGraw-Hill Inc.
Larson, E.D. and Nilsson, L.J., 1991, “Electricity Use and Efficiency in Pumping and Air Handling
Systems, ASHRAE Transactions, pgs. 363-377.
McQuiston and Parker, 1994, Heating Ventilating and Air Conditioning, John Wiley and Sons, Inc.
McQuiston, F., Parker, J. and Spitler, J., 2000, Heating, Ventilating and Air Conditioning: Analysis
and Design, , John Wiley and Sons, Inc.
Mott, 2000, Applied Fluid Mechanics, Prentice Hall, Inc.
Nadel, S., Shepard, M., Greenberg, S., Katz, G., and Almeida, A., 1991, “Energy Efficient
Motor Systems”, American Counsel for an Energy Efficient Economy, Washington D.C.
Tutterow, Energy Efficient Fan Systems, Industrial Energy Technology Conference, Houston, TX.
29