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Comparison of different fan control strategies on a variable air volume systems through simulations and experiments
Accepted Manuscript Comparison of Different Fan Control Strategies on a Variable air volume systmes Through simulations and experiments Gyujin Shim, Li Song, Ph.D., P.E Gang Wang, Ph.D., P.E. PII:
S0360-1323(13)00317-X
DOI:
10.1016/j.buildenv.2013.11.003
Reference:
BAE 3554
To appear in:
Building and Environment
Received Date: 20 August 2013 Revised Date:
1 November 2013
Accepted Date: 5 November 2013
Please cite this article as: Shim G, Song L, Wang G, Comparison of Different Fan Control Strategies on a Variable air volume systmes Through simulations and experiments, Building and Environment (2013), doi: 10.1016/j.buildenv.2013.11.003. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
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• Title:
Comparison of Different Fan Control Strategies on a Variable air volume systmes Through simulations and experiments.
1. Gyujin Shim EL 128, 200 Felgar st., Norman OK73019, U.S.A [email protected] 2. Li Song, Ph.D., P.E FH212, 865 Asp, Norman OK73019, U.S.A [email protected]
3. Gang Wang, Ph.D., P.E.
• Corresponding author.
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Li Song Phone: 1-405.325.1714; Fax: 1-405.325. FH212, 865 Asp, Norman OK73019, U.S.A [email protected]
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• Present/permanent address. The same as above.
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1251 Memorial Drive, McArthur Engineering Building, Rm. 319 [email protected]
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• Author names and affiliations:
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COMPARISON OF DIFFERENT FAN CONTROL STRATEGIES ON VARIABLE AIR VOLUME SYSTEMS THROUGH SIMULATIONS AND EXPERIMENTS
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ABSTRACT
Through a variable frequency drive (VFD), a variable air volume air handling system can significantly reduce supply fan power under partial load conditions. Typically, the VFD on a supply fan motor is modulated to maintain a supply air duct static pressure set point. The static pressure set point can be either constant or dynamically reset in
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response to different load conditions. In this paper, comprehensive mathematical models are established to describe the performance of a VFD-motor-fan system under three different static pressure control strategies: constant static
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pressure set point, static pressure reset by total airflow rate, and static pressure reset by highest zone demand. The total input power to the system, including the energy imparted into the air as well as the energy losses from the VFD, motor, and fan, are simulated and compared among the different static pressure reset strategies with different minimum airflow ratios. The simulation results show that more than 50% of electrical power savings can be realized by static pressure reset when the diversity of the zone thermal loads is moderate. In addition, although a lower
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minimum airflow ratio can result in more power savings, there were no significant power saving by reducing the minimum airflow ratio from 30% to 10% due to relatively high energy losses of the fan, motor and VFD at low load and speed conditions. Finally, experiments were carried out to demonstrate the performance comparison of three
INTRODUCTION
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different control strategies.
A variable air volume (VAV) system is designed to control space air temperature by varying the amount of
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supply air rather than the supply air temperature [1]. The VAV system modulates airflow rate in response to dynamic thermal load changes and thus reduces the fan operating cost, mostly during partial load conditions. A supply fan, a primary component of a VAV system, accounts for significant electricity use in buildings. Therefore, it is necessary to optimize the supply fan operation to save fan electrical power and associated cooling in response to reduced dissipated heat. A closed-loop proportional-integral (PI) control that uses measured duct static pressure at a selected set point in an air distribution system is the most common method of controlling a supply fan that has a variable frequency drive (VFD).
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The static pressure set point is usually a constant, typically 1.5 inch of water (374 Pa), determined by a system design and verified through a system balancing practice during system commissioning. It is essential to place a static pressure sensor properly to maintain satisfactory pressure throughout an air distribution system. A static pressure
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sensor is generally located two-thirds downstream in a main supply air duct. This method is fairly common in many existing systems even though it is no longer recommended by ASHRAE. According to the ASHRAE Application Handbook 2012, performance is satisfactory when the sensor is located on the main duct at 75% to 100% of the distance from the first branch to the most remote terminal [2]. While a fan is modulated to maintain a constant static
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pressure, zone airflow rate is regulated by partially closing VAV terminal dampers during partial load conditions. This operation shows that there is an opportunity to achieve fan power savings by reducing the static pressure set
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point under partial load conditions. Research and practices have focused on static pressure reset strategy development and implementation in order to reduce fan power consumption. In general, these static pressure reset strategies can be categorized into two types: 1) static pressure reset by AHU feedback, and 2) static pressure reset by zone level feedback. For AHU-level feedback control, Liu presented a static pressure reset based on fan speed and demonstrated about 68~75% annual fan power savings in 1998 [3]. Liu (2008) also introduced a reset strategy based on total supply airflow ratio, which is measured by a fan airflow station [4]. For zone level feedback controls,
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Hartman (1989) presented a terminal regulated air volume (TRAV) control strategy in which a supply fan is modulated to meet VAV terminal airflow requirements instead of a static pressure set point [5]. Englander and Norford (1992) proposed two methods of regulating static pressure using a primary flow error signal from one or
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more zones [6]. Warren and Norford (1993) presented 42% fan power savings by resetting static pressure based on the number of starving VAV terminal boxes [7]. Static pressure reset using the actual damper position has been
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experimentally investigated by Rob Moult [8] and numerically by Wang [9]. ASHRAE standard 90.1-2010 [10] recommends that the static pressure set point should be reset for VAV systems equipped with direct digital controls (DDC) at a zone level, which is based on the worst zone requiring the highest pressure.
In general, pressure reset
by AHU feedback is simple but may create starving zones. On the other hand, pressure reset by zone level feedback such as box damper positions is complex and creates exact pressure required by the system, but dynamic damper responses may cause fluctuations of the duct network and the faulty damper position may cause energy waste. The thermal stability of variable pressure control was investigated and evaluated for a variable flow air-conditioning water system by Zhao et. al [11-12].
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Given so many proposed static pressure reset strategies in practice, there is a need to quantify the benefits and penalties and analytically compare the VFD-motor-fan system energy performance and terminal box performance among different strategies. It is important to analyze the total system input power because the efficiency losses of However, most cited
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the VFD, motor and fan at lower load conditions might negate the fan power savings.
research only focuses on mechanical shaft power savings from the fan itself and have neglected the efficiency variations of the motor and VFD. In some cases, the efficiency of fan variations is simply calculated using the cubic relationship of power with airflow rate [13], which is only true for aerodynamically similar fans or equivalent fan
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operation points [14]. Mei and Levermore adopted neural network model to simulate a fan static pressure control in a laboratory VAV test rig [15]. The study was focused on dynamic responses of the system, while the energy
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performance of such a fan system was not discussed. Meanwhile, the most comprehensive building system simulation model, EnergyPlus, uses a design fan performance regression curve plus constant inputs for motor and fan efficiencies to simulate energy consumption. Because fan energy performance is significantly different under different control strategies, a design fan energy performance curve is not sufficient to simulate and compare different system performances under different control strategies [16]. The study presented in this paper is to establish analytical system energy models using fan operation characteristics along with motor and VFD efficiency variations
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under partial speed conditions. The results of the analytical model simulations in this study can be used to compare and analyze the causes of different electrical power consumption rates under different control strategies. In addition, the proposed analytical models can be used to generate different system performance regression curves for different
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operation scenarios, which can be used as input curves to EnergyPlus to precisely simulate the energy performance of a VFD-motor-fan system under different control strategies [17].
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In this paper, analytical VFD-motor-fan operation models that include motor and VFD efficiency degradations under partial speed conditions are introduced first. Second, the performance of one constant static pressure and two static pressure reset strategies is simulated, analyzed, and then compared in order to evaluate the impacts on system input power and terminal box performance. Finally, the comparison of the three different control strategies is demonstrated by experiments conducted on a lab AHU.
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NOMENCLATURE Fan head, Pa or inch of water
N
Power exponent factor
Pd
Duct static pressure at design condition, Pa or inch of water
Pmin
Minimum duct static pressure, Pa or inch of water
Preset
Reset duct static pressure, Pa or inch of water
Q
Airflow rate, m3/s or CFM
S
System resistance coefficient [inch.wg / (ft3/min)2 or Pa/(m3/s)]
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H
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System resistance coefficient from the supply fan to Point III [inch.wg / (ft3/min)2 or Pa/(m3/s)] System resistance coefficient from Point III to Point II [inch.wg / (ft3/min)2 or Pa/(m3/s)]
Winput
System input power, kW
W
Power conveyed to air, kW
ž
Efficiency Relative fan speed
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System resistance coefficient from Point II to Point I [inch.wg / (ft3/min)2 or Pa/(m3/s)]
Pressure loss from the AHU to Point III [inch.wg or Pa]
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Pressure loss from Point III to Point II [inch.wg or Pa]
MODELS
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Pressure loss from Point II to Point I [inch.wg or Pa]
A fan is an air pump that creates a pressure difference and causes airflow [14]. Figure 1 shows a system with a variable frequency drive (VFD), a motor and a fan. The system input electrical power to the VFD-motor-fan system needs to include the energy for delivering the pressure increase, i.e., the fan head (H) required for the amount of airflow rate (Q) as shown in Equation (1) and compensating for efficiency losses from the fan (žfan), motor (žmotor) and VFD (žVFD). The system input power to the VFD-motor-fan system is summarized in Equation (2). Figure 1. Electrical configuration of a fan system.
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,
(1)
W���������� = H · Q⁄(������� · ����������� · �VFD )
is the power conveyed to the air stream; Winput is the system input power to the VFD-motor-fan system
[kW]; H is the fan head [inch.wg or Pa]; Q is the airflow rate [CFM or m3/s]; and fan, motor and VFD efficiency [%], respectively.
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where
(2)
are the
As indicated in Equation (2), in order to compare system input power under different static pressure reset
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strategies, the fan airflow rate, fan head, and fan, motor and VFD efficiencies need to be obtained. Fan airflow rate
For a VAV system, fan air flow rate is determined by zone thermal loads and minimum airflow requirements.
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Thermal load is a heat rate which must be removed from or added to the conditioned space in order to maintain the design indoor condition. The thermal load is determined by outdoor air temperature, space function and occupancy schedules. At design cooling load conditions, 100% design airflow rate is needed.
As the cooling load decreases at
partial load conditions, the airflow rate decreases until it reaches a required minimum airflow rate. For VAV systems, the airflow rate can be reduced according to the load variations to conserve fan power and possible reheat energy.
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However, a minimum airflow rate (MAR) is usually maintained for indoor ventilation purposes during low cooling and heating load conditions. The minimum airflow rate is typically 30% for most buildings specified by ASHRAE [10]. When the load ratio is higher than the MAR, the airflow ratio follows the load ratio. On the other hand, when
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the load ratio is lower than the MAR or becomes negative in heating seasons, the airflow ratio is usually kept at the MAR. Therefore, the airflow rate is a given input as long as the cooling load and minimum airflow requirement are
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known. System Pressure Loss
The fan head (H) compensates for total pressure losses through an entire air distribution system. The system pressure loss (
is proportional to the total system resistance coefficients (S) under fixed damper positions and
the airflow rate squared (Q2) [18]. The airflow rate (Q) is a given input. The system resistance coefficient is determined by the following two factors:
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1) Friction pressure losses and local pressure drops caused by bends or other fittings on ducts. The system resistance coefficients of this type of pressure losses are considered constant in this study because wear on inner duct walls and dust accumulation are assumed negligible. The system resistant coefficients caused by damper
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2) Changes in damper positions on VAV terminals.
position changes vary with different fan operation strategies, and thus they are considered as variables. In the following section, the system pressure loss calculation is introduced according to the three different fan control strategies mentioned in the Introduction. Scenario 1 uses a constant static pressure set point, serving as a
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base case. Scenario 2 uses a total airflow rate to reset the static pressure set point, representing an AHU feedback control strategy. Scenario 3 uses the worst zone requiring the most pressure to reset the duct static pressure, Figure 2 shows a VAV system with three zones represented by three
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representing a zone feedback control strategy.
VAV terminals and a supply fan, where the duct static pressure sensor is located 3/4 downstream from the fan. The duct static pressure set point is determined in a way that the pressure loss of the entire system can be compensated for by the fan, when the fan operates to maintain the duct static pressure set point. Typically, a duct static pressure (P) set point is used for comparison with a measured duct static pressure for fan speed modulation.
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Figure 2. Schematic of a VAV system.
Scenario 1. Constant duct static pressure set point
Scenario 1 is considered as a base case where the static pressure is maintained at a constant throughout different load conditions. Thus, to satisfy the worst scenario, the maximum duct static pressure required by design load, Pd, is
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maintained constantly under partial load conditions. The velocity head at the duct static pressure sensor is lumped into the main duct pressure loss calculation (SI). Under partial load conditions, the system pressure loss is defined by
where
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Equation (3).
(3)
= Pressure loss from the AHU to Point III in Figure 2 [inch.wg or Pa]
= Pressure loss from Point III to Point II in Figure 2 [inch.wg or Pa] = Pressure loss from Point II to Point I in Figure 2 [inch.wg or Pa]
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= System resistance coefficient from the supply fan to Point III in Figure 2 [inch.wg / (ft3/min)2 or Pa/(m3/s)] = System resistance coefficient from Point III to Point II in Figure 2 [inch.wg / (ft3/min)2 or Pa/(m3/s)]
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= System resistance coefficient from Point II to Point I in Figure 2 [inch.wg / (ft3/min)2 or Pa/(m3/s)] As discussed earlier, the resistance coefficients on the main ducts (SIII, SII, SI,) are considered constant because there are no moving parts on the main duct work. Zone resistance coefficients (S3, S2, S1,) vary with terminal dampers’ movements. With reduced airflow under partial space thermal loads, the damper is closed to increase S1 in
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order to maintain the constant pressure loss at Pd. Scenario 2. Duct static pressure reset by total airflow rate in an AHU
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In Scenario 2, the duct static pressure is reset based on the total airflow rate squared, which is determined by the sum of the required airflow rate at each zone. The system pressure drop under partial load conditions can be expressed by Equation (4).
(4)
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Ideally, Preset is reset by the total airflow rate of the AHU,
(5)
However, in actual operations, a safety factor is usually added to calculate Preset to reduce the possibility of
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starving zones, which are potentially caused by diverse zone load profiles for different zones.
,
(6)
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where Pmin is the minimum setting of the duct static pressure as a safety factor. However, a starving zone can still occur if the zone load variation is beyond the safety factor. To determine a starving zone, the pressure provided at each zone by the fan operating to maintain the duct static pressure set point needs to be compared with the pressure required by each individual zone load. Since the terminal box dampers are fully open to maintain the required design airflow rate under the design conditions after system balancing, a starving zone occurs where the terminal box is not able to provide the needed airflow rate with a fully open damper.
If the
provided pressure (Pi,prov), which is calculated from the duct static pressure set point defined in Equation (6), is lower than the needed pressure (Pi,req), which is calculated based on the required zone supply airflow rate with a fully open 7
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zone damper, a starving zone is indicated. In this case, the zone damper is maintained at the maximum open position (100%) and an actual airflow rate is calculated by using the provided pressure ( ), as shown in Equation (7).
, (7) where subscript “i” represents different zones, i.e., 1, 2 and 3 in this study.
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resistance coefficient at a 100% damper open position (
and the zone
Scenario 3. Duct static pressure reset by the worst zone that requires the highest pressure
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In Scenario 3, the duct static pressure is reset to maintain at least one zone damper nearly 100% open. This is usually done by using the terminal box damper position to reset the duct static pressure set point. Compared with
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Scenario 2, which ignores the load diversity of each zone, this method ensures all the zones have enough airflow rates. However, a stuck terminal box damper may over-set the duct static pressure set point and result in excessive fan power, and dynamic damper responses may cause fluctuations of the fan speed controls. To identify the worst zone that needs the highest pressure under partial conditions, the pressure losses at different branches need to be compared and the one with the highest pressure loss becomes the worst branch. The system pressure loss is
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calculated by the equations below:
(8)
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where
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Fan speed
Fan head (H) is required to compensate for the total pressure losses through an entire air distribution system. The fan head at the 100% rated fan speed is given by a design fan curve, which can be regressed by a 2nd-order polynomial equation as shown in Equation (9).
where
(9)
are fan curve coefficients under the 100% rated fan speed.
Fan speed (ω) is one of the factors that determines motor efficiency (žmotor). To calculate fan speed (ω), the fan speed ratio (
) is expressed by the fan affinity laws in Equation (10). 8
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ω = Q⁄Qd 2 ω = H⁄Hd
(10)
where subscript “d” represents 100% of rated fan speed.
(11)
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By integrating Equations (9) and (10), fan heads at partial load conditions can be formulated by Equation (11).
The fan speed at partial load conditions can be obtained to balance the system pressure loss with the fan head defined by Equation (11), as shown in Equation (12).
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(12)
Fan efficiency
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Fan efficiencies (žfan) can be determined using manufacturers’ fan efficiency curves. Figure 3 shows a fan head curve provided by manufacturers, along with a typical fan efficiency curve for a forward-curved blade fan. Typically, the fan efficiency performance curve is given under the rated fan speed. However, fan speed through a VFD varies under partial load conditions. Thus, it is essential to obtain a general fan efficiency curve that is independent of fan speed. According to fan performance characteristics (fan laws), fan efficiency is a function of
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total system resistance coefficient (S), as shown in Equation (13). The S is defined in Equation (14), i.e., once the total S is fixed, the fan efficiency is fixed regardless of the fan speed. ������� = �� ������(S)
(13)
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where
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(14)
Thus, the fan efficiency curve at the rated fan speed in Figure 3 can be converted into Figure 4, which can be used to obtain fan efficiencies for any fan speeds as long as the normalized system resistance coefficient is given. The normalized system resistance coefficient is the ratio of the system resistance coefficient at any fan operation conditions over the system resistance coefficient at the design condition, as shown by the filled circle in Figure 3. Figure 3. Typical fan curves for forward blade fan. Figure 4. Fan efficiency as a function of normalized S.
Motor efficiency 9
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Motor efficiency (žmotor) is related to motor load and power frequency, as expressed in Equation (15), where motor load is determined by the fan shaft work and its efficiency. The power frequency corresponds to the fan speed ratio (ω), which can be obtained through fan operation models or Equation (12). Recent motor efficiency curves
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published by motor manufacturers only indicate motor efficiency changes under rated speeds with variations of motor load. Thus, it requires additional modeling and simulation to generate motor efficiency under variable speeds and partial loads. The results of one 3hp (2.237kW) motor with a linear voltage-frequency relationship are presented in Figure 5, which will be used in the later simulation section. The study was done by Wang et al. [19-20]. More in-
����������� = �� ) ����� �����(H · Q⁄������� , ω
(15)
is the fan speed ratio to indicate the motor frequency [rpm].
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where
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depth analysis can be found in the references, which are not covered in the paper.
Figure 5. Motor efficiency of the 3hp (2.23kW) motor under different frequencies.
VFD efficiency
The VFD controls the rotational speed of a motor by adjusting the frequency and voltage of the power supplied to an AC motor. The efficiency loss of VFDs should be considered when calculating power consumption since the
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VFD efficiency decreases as the motor load decreases. When the motor load is less than 20%, a dramatic loss of efficiency in the VFD can be observed. The VFD efficiency loss is more pronounced with smaller horsepower motors. Figure 6 shows the VFD efficiency curve for a 3 hp (2.23kW) motor [14]. In addition, VFD efficiency is expressed in Equation (16).
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Figure 6. VFD efficiency of 3hp motor.
(16)
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������� = �� ������(H · Q⁄(������� · �����������))
Energy input to the VFD-motor-fan system As a result, the total system input power to a VFD-motor-fan can be obtained by the integration of Equations (2), (14), (15), and (16), shown in Equation (17) and which requires only one independent input variable: the fan airflow rate determined by the load and minimum airflow setting.
The fan head required to compensate for the
total pressure loss can be determined by the airflow rate and the different fan control strategies. W���������� =
ffan
(H⁄Q2 )
H·Q · fmotor (H ∙ Q⁄�������, ω) · fVFD (H · Q⁄(������� · �����������))
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(17)
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SIMULATIONS To visualize the effect of the different duct static pressure controls, simulations were conducted based on the single duct VAV system presented in Figure 2. A supply air fan with 3 hp (2.23 kW) motor and 2,500 CFM (1,180
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L/s) airflow rate is selected in the simulation. The fan, motor, and VFD curves presented in the Models section are used in the simulation. In addition, a set of assumptions is made for the simulation, such as zone load profiles, design pressure losses in the VAV system, and minimum airflow rate settings for VAV systems. All three static pressure reset scenarios introduced previously are simulated in this section. The safety factor introduced in Equation
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(6) for Scenario 2 is taken as 20% of the design static pressure. Simulation conditions Zone load profiles.
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1.
Since the fan airflow rate is determined by the zone loads, it is necessary to define the zone load profiles for all three different zones. To simulate the impacts of different zone load variations, it is assumed that each zone consists of different combinations of both internal and envelope loads. A normalized envelope load profile (load ratio over the maximum design load) is defined as a linear function of outside air temperature, 100% at 102 oF (38.8 oC) outside air design temperature, and 0% at 67 oF (19.4 oC) outside air temperature. In contrast to envelope load, a
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normalized internal load is assumed to be consistent in the simulation since it is related to people, lighting, and equipment. The internal load variations based on occupant changes are ignored in this study. To simulate a typical VAV system, Zone A is assumed with 30% internal load and 70% envelop load, Zone B
2.
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with 50% of both internal load and envelop load, and Zone C with 70% internal load and 30% design envelop load. Design pressure losses in the VAV system.
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Information on design pressure losses in a VAV system is critical for fan operations. Based on AHU air distribution system design principles, in the simulation it is assumed that at the design condition, the pressure loss distributions on the main ducts are ∆PI: ∆PII: ∆PIII= 4:1:1. In addition, it is assumed that the ratio of the total main duct pressure loss over the most remote terminal pressure loss is 3:1 at the design condition. Thus, the ratio of resistance coefficients of individual zone is S1:S2:S3=2:3:4 with all terminal box dampers at a fully open position. 3.
Minimum airflow rate for the VAV system.
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The MAR set point has direct impact on the fan power consumption. In order to understand the impact of different MAR set points on the system performance, three different MAR set points of 50%, 30% and 10% are also defined in the simulations to compare different fan operations for all three scenarios.
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Simulation results Fan operation characteristics under the three scenarios at 30% of MAR are compared in Figure 7 with respect to airflow rate ratio.
The reference curve in Figure 7 represents a system curve under the design condition where all
the terminal dampers are always fully open and the lowest system resistance coefficient occurs during the cooling Figure 7 shows that the fan operation points in Scenario 1 are higher than those in the other two
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load variation.
scenarios. The differences increase as airflow rate ratio decreases. More fan head is provided in Scenario 1 and
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consequently causes terminal box dampers to close to maintain an appropriate amount of supply air flow, leading to higher pressure losses in the air distribution system. More system input power is also consumed. Significant improvement can be observed in Scenarios 2 and 3. It is observed that the operation curves of Scenarios 2 and 3 intersect at 0.39 of airflow rate ratio. To the left of this point, the fan head in Scenario 2 becomes higher than that in Scenario 3. This is due to the 20% safety factor applied in Scenario 2. This indicates that over-pressurization occurs in Scenario 2 when the safety factor kicks in. In contrast, to the right of this point, the fan head in Scenario 2 is
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lower than that in Scenario 3, indicating that starving zones do exist under Scenario 2 operations. Figure 7. Fan operations characteristics w.r.t. airflow ratio.
In order to evaluate the VFD-motor-fan system performance, the VFD-motor-fan power exponent factor is
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introduced as an index. Ideally, if the air distribution system resistance coefficient under a partial load condition is maintained as low as the design system resistance coefficient as shown by the reference line in Figure 7, the fan
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power exponent is 3, as expressed in Equation (18) [21]. Thus, the required fan power under partial load conditions is proportional to the cube of the airflow rate, which sets the minimum fan power consumption for the required airflow rate.
(18)
However, in actual operations, the VFD-motor-fan power exponent factor (N) is usually less than 3 due to two facts: 1) motor and VFD efficiency loss; and 2) fan efficiency changes due to a higher resistant coefficient of an actual system than that in design conditions. In this study, the VFD-motor-fan power exponent factor, as defined in
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Equation (19), is used as an energy performance index to compare different VFD-motor-fan operations among the different scenarios. The smaller the exponent factor is, the less efficient the VFD-motor-fan system is.
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(19) The system input power, power exponent factors, and system resistance coefficient (S) at 30% of MAR are compared in Figures 8(a) to 8(b) with respect to the supply airflow ratio. Similar to the fan operation points in Figure 7, the system input power in Scenario 1 is higher than that in the other two scenarios in Figure 8(a). The
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system input power in Scenarios 2 and 3 has similar results as the fan head curves. More than 50% of system input power savings can be observed in Scenarios 2 and 3 when the total design air flow rate is about 30% to 50% of the design airflow rate, which occurs most of the time in an AHU operation. It is also noticed that the system input
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power in Scenario 3 is higher than that in Scenario 2 when the airflow ratio is higher than 40% due to un-uniformed zone load profiles. In another words, when the total airflow ratio is lower than 40%, the system input power in Scenario 3 is lower than that in Scenario 2 because of the safety factor designed for Scenario 2. In Figure 8(b), system power exponent factors (N) are presented for three scenarios. The system power exponent factors of Scenarios 2 and 3 are higher than that in Scenario 1, which indicates that there are more efficient operations in
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Scenarios 2 and 3.
Figure 8. Operation comparisons w.r.t total airflow ratio for three scenarios.
In Figure 9(a), the fan operations characteristics of Scenario 2 are compared with different MARs ranging from
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10%, 30% to 50%. The fan operation curves overlap with each other when the total airflow ratio ranges from 1 to 0.64 regardless of MARs, because the MAR has no impact on the fan operations at high load conditions. When the
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total airflow ratio decreases to below 0.64, the fan operation curves with 10% MAR have similar fan heads to those with 50% and 30% MARs, although the fan heads with 10% MAR are slightly lower than those with 50% and 30% MARs. This is because more airflow rate is needed when the MAR is larger for higher MARs, so consequently the fan head is higher.
However, the fan operation curves with 10% MAR have higher fan heads than those with 30%
MAR when the airflow ratio is lower than 0.35. By allowing a lower airflow rate using 10% MAR, the airflow distribution among three zones is more diverse than the airflow distribution with 30% MAR; hence, individual terminal box damper regulation is needed the most for the 10% MAR operation. Increased system resistance and fan
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head negate the fan head increases by the increased total airflow rate. The results of Scenarios 1 and 3 are not presented in this article due to similar patterns that were observed as the pattern in Scenario 2. As shown in Figure 9(b), the VFD-motor-fan system input power consumption in Scenario 2 was compared
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with different MARs with respect to the outdoor air temperature, which indicates the load changes based on the simulation assumptions. A similar pattern was observed that the system input power overlaps with each other when the outdoor air is higher than 77°F (25°C), equivalent to 0.64 airflow ratio. When the outdoor air temperature decreases, the system input power is generally reduced by using lower MARs, except that no significant power
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reduction is observed when MAR is reduced from 30% to 10%. This is due to the extreme high efficiency losses of the motor and VFDs at low motor load and speed conditions. Thus, it can be an effective way of reducing system
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input power when a moderate MAR reduction is in effect, such as from 50% to 30%. However, no significant system input power improvement is observed when the MAR is reduced from 30% to 10%. Since a lower MAR usually causes poorer air ventilation and circulation in rooms, caution needs to be taken to determine an appropriate MAR.
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Figure 9. Operation comparison for different MARs.
All the simulation results presented in Figures 7-9 are based on the zone load distribution of 70% envelope load in Zone A, 50% envelop load in Zone B, and 30% envelop load in Zone C. To investigate whether the different load distribution layouts impact the results, another set of simulations was conducted with 30% envelop load for Zone A
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and 70% envelop load for Zone C, while Zone B was kept at the same 50% envelop load. The simulation results showed similar performance. Therefore, the location impact was found to be negligible as long as the loads are
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maintained the same for an AHU. EXPERIMENTS
In order to experimentally compare the energy performance as well as the thermal satisfaction under different fan control scenarios, three different static pressure reset scenarios were also implemented in an actual AHU on a university campus, which has a supply air fan with the same specifications as described in the simulations. The difference is that the actual AHU serves seven terminal boxes, i.e., seven thermal zones including a lobby, faculty and graduate student offices, and two rest rooms. The floor plan is shown in Figure 10. The total conditioned space
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is about 2300 square feet (214m2).
The static pressure sensor is located at about ¾ downstream of the supply air
fan. The experiments were carried out in August and September 2012.
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Figure 10. Serving zone layout of the experimental unit.
The measured fan power consumption, fan head and airflow rate comparisons are shown in Figures 11 and 12. The system input power among three different scenarios is compared in Figure 11. Scenario 1 shows the consistently highest system input power, while Scenario 2 shows the consistently lowest system input power under a similar
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range of airflow rate. Compared with Scenario 1, Scenario 2 saves about 60% system input power at a lower airflow range and 35% of system input power at a higher airflow range. An interesting finding is that the system input power
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in Scenario 3 spreads out between Scenarios 1 and 2, which is different from the simulation results. By comparing measured static pressure, shown in Figure 12(a), Scenario 3 maintained much higher duct static pressure than Scenario 2 along with expected oscillations. This is due to the more severe load diversity among the thermal zones than the one assumed in the simulations. Even though Scenario 2 shows much better energy performance, starved thermal zones occur as indicated in Figure 12(b), where the summation of the total supply airflow rate from all the
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terminal boxes is around 25% lower than the summation of the supply airflow set points. The outdoor air temperature is shown in the same figure using the label on the right side. The results indicate that some thermal zones may suffer thermal problems. On the other hand, the measured airflow rate is well maintained at the airflow
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rate set points by Scenario 3.
Figure 11. Measured system input power comparison among three different scenarios.
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Figure 12. Measured performance comparison among three different scenarios.
CONCLUSION
A comprehensive mathematical model to calculate VFD-motor-fan input power was established by including supply airflow rate, pressure losses, fan efficiency, motor efficiency, and VFD efficiency as inputs in this paper. Simulations were conducted to compare fan operations, system input power, power exponent factors, and system resistance coefficients under three different fan control scenarios: 1) using a constant static pressure set point; 2) resetting the static pressure using total airflow rate (feedback from AHUs); and 3) resetting static pressure using the
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worst zone (feedback from zones). The simulation results show that more than 50% of fan power consumption can be observed in Scenarios 2 and 3 under partial load conditions. In addition, the impact of minimum airflow ratio (MAR) on the fan power energy performance was also investigated. The results show that moderate MAR
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reductions can result in a reasonable amount of power savings. It is also concluded that for the simulation case (the 3hp fan), no significant system input power savings was observed by reducing MAR from 30% to 10% due to the increased efficiency losses from the fan, motor and VFD at low load and speed conditions. Experiments has also approved that up to 60% of system input power savings can be obtained in Scenario 2 (the pressure reset by the total
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airflow rate) compared with Scenario 1. However, the performance of Scenario 3 heavily relies on the distribution of the zone loads. Unbalanced zone loads significantly reduced fan power savings compared with the simulation
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REFERENCES
[1] ASHRAE. HVAC systems and equipment. Chapter 4. Air handling and Distribution. Atlanta : American Society of Heating, Refrigerating and Air-conditioning Engineers, Inc; 2008, p.4.11 [2] ASHRAE.
HVAC application. Chapter 46. Design and application of controls. Atlanta : American Society of
[3] Liu
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Heating, Refrigerating and Air-conditioning Engineers, Inc; 2011, p.46.7-46.8 M, Zhu Y, Powell T, Claridge D. E. System Optimization Saves $195,000/yr. in a New Medical Facility.
Proceedings of the 6th National Conference on Building Commissioning 1998; 14.21-14.2.11
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[4] Liu G, Liu M. Supply Fan Control Methods for VAV Systems Using a Fan Airflow Station. ASHRAE Transactions 2008; 114(1):451-457
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[5] Hartman T, TRAV. A new HVAC concept. Heating, piping and air conditioning 1989; 61(7):69-73 [6] Englander S, Norford LK, Saving fan energy in VAV systems-Part 2: Supply fan control for static pressure minimization using DDC zone feedback. ASHRAE Transactions 1992; 98(1):19-32 [7] Warren M, Norford LK, Integrating VAV zone requirement with supply fan operation. ASHRAE Journal 1993; 35(4):43-46
[8] Moult R, Control strategies for VAV Air Handling Units-Part2. Air Conditioning and Refrigeration Journal. 1999; http://www.ishrae.in/journals/1999apr/article03.html
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[9] Wang S, Dynamic simulation of building VAV air-conditioning system and evaluation of EMCS on-line control strategies, Building and Environment 34 (1999): 681-705. [10] ANSI/ASHRAE/IESNA Standard 90.1-2010. Energy Standard for Buildings Except Low-Rise Residential
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Buildings. Atlanta : American Society of Heating, Refrigerating and Air-conditioning Engineers, Inc; 2010 [11] Zhao T, Zhang J, Ma L, Experimental analysis of thermal stability of the pressure control method for a variable flow air-conditioning water system, Building and Environment 70 (2013): 1-9.
[12] Zhao T, Zhang J, Ma L, On-line optimization control method based on extreme value analysis for parallel
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variable-frequency hydraulic pumps in central air-conditioning systems, Building and Environment 47 (2012): 330-338.
and Environment 35 (2000): 471-487.
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[13] Wang S, Jin X, Model-based optimal control of VAV air-conditioning system using genetic algorithm, Building
[14] ASHRAE. HVAC Systems and equipment. Chapter 20.1. Fans.
Atlanta : American Society of Heating,
Refrigerating and Air-conditioning Engineers, Inc; 2008, p.20.
[15] Mei L, Levermore G.J., Simulation and validation of a VAV system with an ANN fan model an a non-linear VAV box model, Building and Environment 37 (2002):277-284.
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[16] Lawrence Berkeley National Laboratory (LBNL), EnergyPlus Engineering Reference – Air system Fans. 2011;704-724 http://apps1.eere.energy.gov/buildings/energyplus/energyplus_documentation.cfm [17] Branesky B, Song L, Air handling unit level fault signature development using EnergyPlus, Proceedings of
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ASME2012 International Mechanical Engineering Congress and Exposition, IMECE2010-87574, Houston TX. [18] ASHRAE. HVAC Systems and Equipment. Chapter 20.4. Fans. .
Atlanta : American Society of Heating,
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Refrigerating and Air-conditioning Engineers, Inc; 2008, p.20 [19] Wang G, Liu M, Claridge DE,. Development of an Energy Meter Using a Pump Flow Station. ASHRAE Transactions 2010;116(2)
[20] Wang G, Song L, Park S, Estimation of Induction Motor Efficiency under Variable Frequencies.
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Transactions, in press
[21] U.S. Department of Energy - Energy Efficiency and Renewable Energy. Adjustable Speed Drive Part-Load Efficiency 2008; http://www1.eere.energy.gov/manufacturing/tech_deployment/pdfs/motor_tip_sheet11.pdf
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A comprehensive mathematical model is built to calculate VFD-motor-fan power. 3 scenarios include no reset, reset by total flow and rest by highest zone demand. The model can generate input curves to EnergyPlus to simulate a VFD-motor-fan system. The simulation results show that more than 50% savings. Experiments show the zone feedback control relies on zone load diversity.
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• • • • •
S0360-1323(13)00317-X
DOI:
10.1016/j.buildenv.2013.11.003
Reference:
BAE 3554
To appear in:
Building and Environment
Received Date: 20 August 2013 Revised Date:
1 November 2013
Accepted Date: 5 November 2013
Please cite this article as: Shim G, Song L, Wang G, Comparison of Different Fan Control Strategies on a Variable air volume systmes Through simulations and experiments, Building and Environment (2013), doi: 10.1016/j.buildenv.2013.11.003. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
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• Title:
Comparison of Different Fan Control Strategies on a Variable air volume systmes Through simulations and experiments.
1. Gyujin Shim EL 128, 200 Felgar st., Norman OK73019, U.S.A [email protected] 2. Li Song, Ph.D., P.E FH212, 865 Asp, Norman OK73019, U.S.A [email protected]
3. Gang Wang, Ph.D., P.E.
• Corresponding author.
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Li Song Phone: 1-405.325.1714; Fax: 1-405.325. FH212, 865 Asp, Norman OK73019, U.S.A [email protected]
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• Present/permanent address. The same as above.
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1251 Memorial Drive, McArthur Engineering Building, Rm. 319 [email protected]
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• Author names and affiliations:
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COMPARISON OF DIFFERENT FAN CONTROL STRATEGIES ON VARIABLE AIR VOLUME SYSTEMS THROUGH SIMULATIONS AND EXPERIMENTS
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ABSTRACT
Through a variable frequency drive (VFD), a variable air volume air handling system can significantly reduce supply fan power under partial load conditions. Typically, the VFD on a supply fan motor is modulated to maintain a supply air duct static pressure set point. The static pressure set point can be either constant or dynamically reset in
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response to different load conditions. In this paper, comprehensive mathematical models are established to describe the performance of a VFD-motor-fan system under three different static pressure control strategies: constant static
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pressure set point, static pressure reset by total airflow rate, and static pressure reset by highest zone demand. The total input power to the system, including the energy imparted into the air as well as the energy losses from the VFD, motor, and fan, are simulated and compared among the different static pressure reset strategies with different minimum airflow ratios. The simulation results show that more than 50% of electrical power savings can be realized by static pressure reset when the diversity of the zone thermal loads is moderate. In addition, although a lower
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minimum airflow ratio can result in more power savings, there were no significant power saving by reducing the minimum airflow ratio from 30% to 10% due to relatively high energy losses of the fan, motor and VFD at low load and speed conditions. Finally, experiments were carried out to demonstrate the performance comparison of three
INTRODUCTION
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different control strategies.
A variable air volume (VAV) system is designed to control space air temperature by varying the amount of
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supply air rather than the supply air temperature [1]. The VAV system modulates airflow rate in response to dynamic thermal load changes and thus reduces the fan operating cost, mostly during partial load conditions. A supply fan, a primary component of a VAV system, accounts for significant electricity use in buildings. Therefore, it is necessary to optimize the supply fan operation to save fan electrical power and associated cooling in response to reduced dissipated heat. A closed-loop proportional-integral (PI) control that uses measured duct static pressure at a selected set point in an air distribution system is the most common method of controlling a supply fan that has a variable frequency drive (VFD).
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The static pressure set point is usually a constant, typically 1.5 inch of water (374 Pa), determined by a system design and verified through a system balancing practice during system commissioning. It is essential to place a static pressure sensor properly to maintain satisfactory pressure throughout an air distribution system. A static pressure
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sensor is generally located two-thirds downstream in a main supply air duct. This method is fairly common in many existing systems even though it is no longer recommended by ASHRAE. According to the ASHRAE Application Handbook 2012, performance is satisfactory when the sensor is located on the main duct at 75% to 100% of the distance from the first branch to the most remote terminal [2]. While a fan is modulated to maintain a constant static
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pressure, zone airflow rate is regulated by partially closing VAV terminal dampers during partial load conditions. This operation shows that there is an opportunity to achieve fan power savings by reducing the static pressure set
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point under partial load conditions. Research and practices have focused on static pressure reset strategy development and implementation in order to reduce fan power consumption. In general, these static pressure reset strategies can be categorized into two types: 1) static pressure reset by AHU feedback, and 2) static pressure reset by zone level feedback. For AHU-level feedback control, Liu presented a static pressure reset based on fan speed and demonstrated about 68~75% annual fan power savings in 1998 [3]. Liu (2008) also introduced a reset strategy based on total supply airflow ratio, which is measured by a fan airflow station [4]. For zone level feedback controls,
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Hartman (1989) presented a terminal regulated air volume (TRAV) control strategy in which a supply fan is modulated to meet VAV terminal airflow requirements instead of a static pressure set point [5]. Englander and Norford (1992) proposed two methods of regulating static pressure using a primary flow error signal from one or
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more zones [6]. Warren and Norford (1993) presented 42% fan power savings by resetting static pressure based on the number of starving VAV terminal boxes [7]. Static pressure reset using the actual damper position has been
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experimentally investigated by Rob Moult [8] and numerically by Wang [9]. ASHRAE standard 90.1-2010 [10] recommends that the static pressure set point should be reset for VAV systems equipped with direct digital controls (DDC) at a zone level, which is based on the worst zone requiring the highest pressure.
In general, pressure reset
by AHU feedback is simple but may create starving zones. On the other hand, pressure reset by zone level feedback such as box damper positions is complex and creates exact pressure required by the system, but dynamic damper responses may cause fluctuations of the duct network and the faulty damper position may cause energy waste. The thermal stability of variable pressure control was investigated and evaluated for a variable flow air-conditioning water system by Zhao et. al [11-12].
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Given so many proposed static pressure reset strategies in practice, there is a need to quantify the benefits and penalties and analytically compare the VFD-motor-fan system energy performance and terminal box performance among different strategies. It is important to analyze the total system input power because the efficiency losses of However, most cited
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the VFD, motor and fan at lower load conditions might negate the fan power savings.
research only focuses on mechanical shaft power savings from the fan itself and have neglected the efficiency variations of the motor and VFD. In some cases, the efficiency of fan variations is simply calculated using the cubic relationship of power with airflow rate [13], which is only true for aerodynamically similar fans or equivalent fan
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operation points [14]. Mei and Levermore adopted neural network model to simulate a fan static pressure control in a laboratory VAV test rig [15]. The study was focused on dynamic responses of the system, while the energy
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performance of such a fan system was not discussed. Meanwhile, the most comprehensive building system simulation model, EnergyPlus, uses a design fan performance regression curve plus constant inputs for motor and fan efficiencies to simulate energy consumption. Because fan energy performance is significantly different under different control strategies, a design fan energy performance curve is not sufficient to simulate and compare different system performances under different control strategies [16]. The study presented in this paper is to establish analytical system energy models using fan operation characteristics along with motor and VFD efficiency variations
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under partial speed conditions. The results of the analytical model simulations in this study can be used to compare and analyze the causes of different electrical power consumption rates under different control strategies. In addition, the proposed analytical models can be used to generate different system performance regression curves for different
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operation scenarios, which can be used as input curves to EnergyPlus to precisely simulate the energy performance of a VFD-motor-fan system under different control strategies [17].
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In this paper, analytical VFD-motor-fan operation models that include motor and VFD efficiency degradations under partial speed conditions are introduced first. Second, the performance of one constant static pressure and two static pressure reset strategies is simulated, analyzed, and then compared in order to evaluate the impacts on system input power and terminal box performance. Finally, the comparison of the three different control strategies is demonstrated by experiments conducted on a lab AHU.
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NOMENCLATURE Fan head, Pa or inch of water
N
Power exponent factor
Pd
Duct static pressure at design condition, Pa or inch of water
Pmin
Minimum duct static pressure, Pa or inch of water
Preset
Reset duct static pressure, Pa or inch of water
Q
Airflow rate, m3/s or CFM
S
System resistance coefficient [inch.wg / (ft3/min)2 or Pa/(m3/s)]
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H
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System resistance coefficient from the supply fan to Point III [inch.wg / (ft3/min)2 or Pa/(m3/s)] System resistance coefficient from Point III to Point II [inch.wg / (ft3/min)2 or Pa/(m3/s)]
Winput
System input power, kW
W
Power conveyed to air, kW
ž
Efficiency Relative fan speed
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System resistance coefficient from Point II to Point I [inch.wg / (ft3/min)2 or Pa/(m3/s)]
Pressure loss from the AHU to Point III [inch.wg or Pa]
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Pressure loss from Point III to Point II [inch.wg or Pa]
MODELS
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Pressure loss from Point II to Point I [inch.wg or Pa]
A fan is an air pump that creates a pressure difference and causes airflow [14]. Figure 1 shows a system with a variable frequency drive (VFD), a motor and a fan. The system input electrical power to the VFD-motor-fan system needs to include the energy for delivering the pressure increase, i.e., the fan head (H) required for the amount of airflow rate (Q) as shown in Equation (1) and compensating for efficiency losses from the fan (žfan), motor (žmotor) and VFD (žVFD). The system input power to the VFD-motor-fan system is summarized in Equation (2). Figure 1. Electrical configuration of a fan system.
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,
(1)
W���������� = H · Q⁄(������� · ����������� · �VFD )
is the power conveyed to the air stream; Winput is the system input power to the VFD-motor-fan system
[kW]; H is the fan head [inch.wg or Pa]; Q is the airflow rate [CFM or m3/s]; and fan, motor and VFD efficiency [%], respectively.
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where
(2)
are the
As indicated in Equation (2), in order to compare system input power under different static pressure reset
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strategies, the fan airflow rate, fan head, and fan, motor and VFD efficiencies need to be obtained. Fan airflow rate
For a VAV system, fan air flow rate is determined by zone thermal loads and minimum airflow requirements.
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Thermal load is a heat rate which must be removed from or added to the conditioned space in order to maintain the design indoor condition. The thermal load is determined by outdoor air temperature, space function and occupancy schedules. At design cooling load conditions, 100% design airflow rate is needed.
As the cooling load decreases at
partial load conditions, the airflow rate decreases until it reaches a required minimum airflow rate. For VAV systems, the airflow rate can be reduced according to the load variations to conserve fan power and possible reheat energy.
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However, a minimum airflow rate (MAR) is usually maintained for indoor ventilation purposes during low cooling and heating load conditions. The minimum airflow rate is typically 30% for most buildings specified by ASHRAE [10]. When the load ratio is higher than the MAR, the airflow ratio follows the load ratio. On the other hand, when
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the load ratio is lower than the MAR or becomes negative in heating seasons, the airflow ratio is usually kept at the MAR. Therefore, the airflow rate is a given input as long as the cooling load and minimum airflow requirement are
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known. System Pressure Loss
The fan head (H) compensates for total pressure losses through an entire air distribution system. The system pressure loss (
is proportional to the total system resistance coefficients (S) under fixed damper positions and
the airflow rate squared (Q2) [18]. The airflow rate (Q) is a given input. The system resistance coefficient is determined by the following two factors:
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1) Friction pressure losses and local pressure drops caused by bends or other fittings on ducts. The system resistance coefficients of this type of pressure losses are considered constant in this study because wear on inner duct walls and dust accumulation are assumed negligible. The system resistant coefficients caused by damper
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2) Changes in damper positions on VAV terminals.
position changes vary with different fan operation strategies, and thus they are considered as variables. In the following section, the system pressure loss calculation is introduced according to the three different fan control strategies mentioned in the Introduction. Scenario 1 uses a constant static pressure set point, serving as a
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base case. Scenario 2 uses a total airflow rate to reset the static pressure set point, representing an AHU feedback control strategy. Scenario 3 uses the worst zone requiring the most pressure to reset the duct static pressure, Figure 2 shows a VAV system with three zones represented by three
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representing a zone feedback control strategy.
VAV terminals and a supply fan, where the duct static pressure sensor is located 3/4 downstream from the fan. The duct static pressure set point is determined in a way that the pressure loss of the entire system can be compensated for by the fan, when the fan operates to maintain the duct static pressure set point. Typically, a duct static pressure (P) set point is used for comparison with a measured duct static pressure for fan speed modulation.
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Figure 2. Schematic of a VAV system.
Scenario 1. Constant duct static pressure set point
Scenario 1 is considered as a base case where the static pressure is maintained at a constant throughout different load conditions. Thus, to satisfy the worst scenario, the maximum duct static pressure required by design load, Pd, is
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maintained constantly under partial load conditions. The velocity head at the duct static pressure sensor is lumped into the main duct pressure loss calculation (SI). Under partial load conditions, the system pressure loss is defined by
where
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Equation (3).
(3)
= Pressure loss from the AHU to Point III in Figure 2 [inch.wg or Pa]
= Pressure loss from Point III to Point II in Figure 2 [inch.wg or Pa] = Pressure loss from Point II to Point I in Figure 2 [inch.wg or Pa]
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= System resistance coefficient from the supply fan to Point III in Figure 2 [inch.wg / (ft3/min)2 or Pa/(m3/s)] = System resistance coefficient from Point III to Point II in Figure 2 [inch.wg / (ft3/min)2 or Pa/(m3/s)]
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= System resistance coefficient from Point II to Point I in Figure 2 [inch.wg / (ft3/min)2 or Pa/(m3/s)] As discussed earlier, the resistance coefficients on the main ducts (SIII, SII, SI,) are considered constant because there are no moving parts on the main duct work. Zone resistance coefficients (S3, S2, S1,) vary with terminal dampers’ movements. With reduced airflow under partial space thermal loads, the damper is closed to increase S1 in
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order to maintain the constant pressure loss at Pd. Scenario 2. Duct static pressure reset by total airflow rate in an AHU
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In Scenario 2, the duct static pressure is reset based on the total airflow rate squared, which is determined by the sum of the required airflow rate at each zone. The system pressure drop under partial load conditions can be expressed by Equation (4).
(4)
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Ideally, Preset is reset by the total airflow rate of the AHU,
(5)
However, in actual operations, a safety factor is usually added to calculate Preset to reduce the possibility of
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starving zones, which are potentially caused by diverse zone load profiles for different zones.
,
(6)
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where Pmin is the minimum setting of the duct static pressure as a safety factor. However, a starving zone can still occur if the zone load variation is beyond the safety factor. To determine a starving zone, the pressure provided at each zone by the fan operating to maintain the duct static pressure set point needs to be compared with the pressure required by each individual zone load. Since the terminal box dampers are fully open to maintain the required design airflow rate under the design conditions after system balancing, a starving zone occurs where the terminal box is not able to provide the needed airflow rate with a fully open damper.
If the
provided pressure (Pi,prov), which is calculated from the duct static pressure set point defined in Equation (6), is lower than the needed pressure (Pi,req), which is calculated based on the required zone supply airflow rate with a fully open 7
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zone damper, a starving zone is indicated. In this case, the zone damper is maintained at the maximum open position (100%) and an actual airflow rate is calculated by using the provided pressure ( ), as shown in Equation (7).
, (7) where subscript “i” represents different zones, i.e., 1, 2 and 3 in this study.
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resistance coefficient at a 100% damper open position (
and the zone
Scenario 3. Duct static pressure reset by the worst zone that requires the highest pressure
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In Scenario 3, the duct static pressure is reset to maintain at least one zone damper nearly 100% open. This is usually done by using the terminal box damper position to reset the duct static pressure set point. Compared with
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Scenario 2, which ignores the load diversity of each zone, this method ensures all the zones have enough airflow rates. However, a stuck terminal box damper may over-set the duct static pressure set point and result in excessive fan power, and dynamic damper responses may cause fluctuations of the fan speed controls. To identify the worst zone that needs the highest pressure under partial conditions, the pressure losses at different branches need to be compared and the one with the highest pressure loss becomes the worst branch. The system pressure loss is
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calculated by the equations below:
(8)
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where
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Fan speed
Fan head (H) is required to compensate for the total pressure losses through an entire air distribution system. The fan head at the 100% rated fan speed is given by a design fan curve, which can be regressed by a 2nd-order polynomial equation as shown in Equation (9).
where
(9)
are fan curve coefficients under the 100% rated fan speed.
Fan speed (ω) is one of the factors that determines motor efficiency (žmotor). To calculate fan speed (ω), the fan speed ratio (
) is expressed by the fan affinity laws in Equation (10). 8
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ω = Q⁄Qd 2 ω = H⁄Hd
(10)
where subscript “d” represents 100% of rated fan speed.
(11)
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By integrating Equations (9) and (10), fan heads at partial load conditions can be formulated by Equation (11).
The fan speed at partial load conditions can be obtained to balance the system pressure loss with the fan head defined by Equation (11), as shown in Equation (12).
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(12)
Fan efficiency
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Fan efficiencies (žfan) can be determined using manufacturers’ fan efficiency curves. Figure 3 shows a fan head curve provided by manufacturers, along with a typical fan efficiency curve for a forward-curved blade fan. Typically, the fan efficiency performance curve is given under the rated fan speed. However, fan speed through a VFD varies under partial load conditions. Thus, it is essential to obtain a general fan efficiency curve that is independent of fan speed. According to fan performance characteristics (fan laws), fan efficiency is a function of
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total system resistance coefficient (S), as shown in Equation (13). The S is defined in Equation (14), i.e., once the total S is fixed, the fan efficiency is fixed regardless of the fan speed. ������� = �� ������(S)
(13)
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where
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(14)
Thus, the fan efficiency curve at the rated fan speed in Figure 3 can be converted into Figure 4, which can be used to obtain fan efficiencies for any fan speeds as long as the normalized system resistance coefficient is given. The normalized system resistance coefficient is the ratio of the system resistance coefficient at any fan operation conditions over the system resistance coefficient at the design condition, as shown by the filled circle in Figure 3. Figure 3. Typical fan curves for forward blade fan. Figure 4. Fan efficiency as a function of normalized S.
Motor efficiency 9
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Motor efficiency (žmotor) is related to motor load and power frequency, as expressed in Equation (15), where motor load is determined by the fan shaft work and its efficiency. The power frequency corresponds to the fan speed ratio (ω), which can be obtained through fan operation models or Equation (12). Recent motor efficiency curves
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published by motor manufacturers only indicate motor efficiency changes under rated speeds with variations of motor load. Thus, it requires additional modeling and simulation to generate motor efficiency under variable speeds and partial loads. The results of one 3hp (2.237kW) motor with a linear voltage-frequency relationship are presented in Figure 5, which will be used in the later simulation section. The study was done by Wang et al. [19-20]. More in-
����������� = �� ) ����� �����(H · Q⁄������� , ω
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is the fan speed ratio to indicate the motor frequency [rpm].
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where
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depth analysis can be found in the references, which are not covered in the paper.
Figure 5. Motor efficiency of the 3hp (2.23kW) motor under different frequencies.
VFD efficiency
The VFD controls the rotational speed of a motor by adjusting the frequency and voltage of the power supplied to an AC motor. The efficiency loss of VFDs should be considered when calculating power consumption since the
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VFD efficiency decreases as the motor load decreases. When the motor load is less than 20%, a dramatic loss of efficiency in the VFD can be observed. The VFD efficiency loss is more pronounced with smaller horsepower motors. Figure 6 shows the VFD efficiency curve for a 3 hp (2.23kW) motor [14]. In addition, VFD efficiency is expressed in Equation (16).
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Figure 6. VFD efficiency of 3hp motor.
(16)
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������� = �� ������(H · Q⁄(������� · �����������))
Energy input to the VFD-motor-fan system As a result, the total system input power to a VFD-motor-fan can be obtained by the integration of Equations (2), (14), (15), and (16), shown in Equation (17) and which requires only one independent input variable: the fan airflow rate determined by the load and minimum airflow setting.
The fan head required to compensate for the
total pressure loss can be determined by the airflow rate and the different fan control strategies. W���������� =
ffan
(H⁄Q2 )
H·Q · fmotor (H ∙ Q⁄�������, ω) · fVFD (H · Q⁄(������� · �����������))
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SIMULATIONS To visualize the effect of the different duct static pressure controls, simulations were conducted based on the single duct VAV system presented in Figure 2. A supply air fan with 3 hp (2.23 kW) motor and 2,500 CFM (1,180
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L/s) airflow rate is selected in the simulation. The fan, motor, and VFD curves presented in the Models section are used in the simulation. In addition, a set of assumptions is made for the simulation, such as zone load profiles, design pressure losses in the VAV system, and minimum airflow rate settings for VAV systems. All three static pressure reset scenarios introduced previously are simulated in this section. The safety factor introduced in Equation
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(6) for Scenario 2 is taken as 20% of the design static pressure. Simulation conditions Zone load profiles.
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1.
Since the fan airflow rate is determined by the zone loads, it is necessary to define the zone load profiles for all three different zones. To simulate the impacts of different zone load variations, it is assumed that each zone consists of different combinations of both internal and envelope loads. A normalized envelope load profile (load ratio over the maximum design load) is defined as a linear function of outside air temperature, 100% at 102 oF (38.8 oC) outside air design temperature, and 0% at 67 oF (19.4 oC) outside air temperature. In contrast to envelope load, a
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normalized internal load is assumed to be consistent in the simulation since it is related to people, lighting, and equipment. The internal load variations based on occupant changes are ignored in this study. To simulate a typical VAV system, Zone A is assumed with 30% internal load and 70% envelop load, Zone B
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with 50% of both internal load and envelop load, and Zone C with 70% internal load and 30% design envelop load. Design pressure losses in the VAV system.
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Information on design pressure losses in a VAV system is critical for fan operations. Based on AHU air distribution system design principles, in the simulation it is assumed that at the design condition, the pressure loss distributions on the main ducts are ∆PI: ∆PII: ∆PIII= 4:1:1. In addition, it is assumed that the ratio of the total main duct pressure loss over the most remote terminal pressure loss is 3:1 at the design condition. Thus, the ratio of resistance coefficients of individual zone is S1:S2:S3=2:3:4 with all terminal box dampers at a fully open position. 3.
Minimum airflow rate for the VAV system.
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The MAR set point has direct impact on the fan power consumption. In order to understand the impact of different MAR set points on the system performance, three different MAR set points of 50%, 30% and 10% are also defined in the simulations to compare different fan operations for all three scenarios.
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Simulation results Fan operation characteristics under the three scenarios at 30% of MAR are compared in Figure 7 with respect to airflow rate ratio.
The reference curve in Figure 7 represents a system curve under the design condition where all
the terminal dampers are always fully open and the lowest system resistance coefficient occurs during the cooling Figure 7 shows that the fan operation points in Scenario 1 are higher than those in the other two
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load variation.
scenarios. The differences increase as airflow rate ratio decreases. More fan head is provided in Scenario 1 and
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consequently causes terminal box dampers to close to maintain an appropriate amount of supply air flow, leading to higher pressure losses in the air distribution system. More system input power is also consumed. Significant improvement can be observed in Scenarios 2 and 3. It is observed that the operation curves of Scenarios 2 and 3 intersect at 0.39 of airflow rate ratio. To the left of this point, the fan head in Scenario 2 becomes higher than that in Scenario 3. This is due to the 20% safety factor applied in Scenario 2. This indicates that over-pressurization occurs in Scenario 2 when the safety factor kicks in. In contrast, to the right of this point, the fan head in Scenario 2 is
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lower than that in Scenario 3, indicating that starving zones do exist under Scenario 2 operations. Figure 7. Fan operations characteristics w.r.t. airflow ratio.
In order to evaluate the VFD-motor-fan system performance, the VFD-motor-fan power exponent factor is
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introduced as an index. Ideally, if the air distribution system resistance coefficient under a partial load condition is maintained as low as the design system resistance coefficient as shown by the reference line in Figure 7, the fan
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power exponent is 3, as expressed in Equation (18) [21]. Thus, the required fan power under partial load conditions is proportional to the cube of the airflow rate, which sets the minimum fan power consumption for the required airflow rate.
(18)
However, in actual operations, the VFD-motor-fan power exponent factor (N) is usually less than 3 due to two facts: 1) motor and VFD efficiency loss; and 2) fan efficiency changes due to a higher resistant coefficient of an actual system than that in design conditions. In this study, the VFD-motor-fan power exponent factor, as defined in
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Equation (19), is used as an energy performance index to compare different VFD-motor-fan operations among the different scenarios. The smaller the exponent factor is, the less efficient the VFD-motor-fan system is.
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(19) The system input power, power exponent factors, and system resistance coefficient (S) at 30% of MAR are compared in Figures 8(a) to 8(b) with respect to the supply airflow ratio. Similar to the fan operation points in Figure 7, the system input power in Scenario 1 is higher than that in the other two scenarios in Figure 8(a). The
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system input power in Scenarios 2 and 3 has similar results as the fan head curves. More than 50% of system input power savings can be observed in Scenarios 2 and 3 when the total design air flow rate is about 30% to 50% of the design airflow rate, which occurs most of the time in an AHU operation. It is also noticed that the system input
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power in Scenario 3 is higher than that in Scenario 2 when the airflow ratio is higher than 40% due to un-uniformed zone load profiles. In another words, when the total airflow ratio is lower than 40%, the system input power in Scenario 3 is lower than that in Scenario 2 because of the safety factor designed for Scenario 2. In Figure 8(b), system power exponent factors (N) are presented for three scenarios. The system power exponent factors of Scenarios 2 and 3 are higher than that in Scenario 1, which indicates that there are more efficient operations in
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Scenarios 2 and 3.
Figure 8. Operation comparisons w.r.t total airflow ratio for three scenarios.
In Figure 9(a), the fan operations characteristics of Scenario 2 are compared with different MARs ranging from
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10%, 30% to 50%. The fan operation curves overlap with each other when the total airflow ratio ranges from 1 to 0.64 regardless of MARs, because the MAR has no impact on the fan operations at high load conditions. When the
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total airflow ratio decreases to below 0.64, the fan operation curves with 10% MAR have similar fan heads to those with 50% and 30% MARs, although the fan heads with 10% MAR are slightly lower than those with 50% and 30% MARs. This is because more airflow rate is needed when the MAR is larger for higher MARs, so consequently the fan head is higher.
However, the fan operation curves with 10% MAR have higher fan heads than those with 30%
MAR when the airflow ratio is lower than 0.35. By allowing a lower airflow rate using 10% MAR, the airflow distribution among three zones is more diverse than the airflow distribution with 30% MAR; hence, individual terminal box damper regulation is needed the most for the 10% MAR operation. Increased system resistance and fan
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head negate the fan head increases by the increased total airflow rate. The results of Scenarios 1 and 3 are not presented in this article due to similar patterns that were observed as the pattern in Scenario 2. As shown in Figure 9(b), the VFD-motor-fan system input power consumption in Scenario 2 was compared
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with different MARs with respect to the outdoor air temperature, which indicates the load changes based on the simulation assumptions. A similar pattern was observed that the system input power overlaps with each other when the outdoor air is higher than 77°F (25°C), equivalent to 0.64 airflow ratio. When the outdoor air temperature decreases, the system input power is generally reduced by using lower MARs, except that no significant power
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reduction is observed when MAR is reduced from 30% to 10%. This is due to the extreme high efficiency losses of the motor and VFDs at low motor load and speed conditions. Thus, it can be an effective way of reducing system
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input power when a moderate MAR reduction is in effect, such as from 50% to 30%. However, no significant system input power improvement is observed when the MAR is reduced from 30% to 10%. Since a lower MAR usually causes poorer air ventilation and circulation in rooms, caution needs to be taken to determine an appropriate MAR.
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Figure 9. Operation comparison for different MARs.
All the simulation results presented in Figures 7-9 are based on the zone load distribution of 70% envelope load in Zone A, 50% envelop load in Zone B, and 30% envelop load in Zone C. To investigate whether the different load distribution layouts impact the results, another set of simulations was conducted with 30% envelop load for Zone A
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and 70% envelop load for Zone C, while Zone B was kept at the same 50% envelop load. The simulation results showed similar performance. Therefore, the location impact was found to be negligible as long as the loads are
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maintained the same for an AHU. EXPERIMENTS
In order to experimentally compare the energy performance as well as the thermal satisfaction under different fan control scenarios, three different static pressure reset scenarios were also implemented in an actual AHU on a university campus, which has a supply air fan with the same specifications as described in the simulations. The difference is that the actual AHU serves seven terminal boxes, i.e., seven thermal zones including a lobby, faculty and graduate student offices, and two rest rooms. The floor plan is shown in Figure 10. The total conditioned space
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is about 2300 square feet (214m2).
The static pressure sensor is located at about ¾ downstream of the supply air
fan. The experiments were carried out in August and September 2012.
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Figure 10. Serving zone layout of the experimental unit.
The measured fan power consumption, fan head and airflow rate comparisons are shown in Figures 11 and 12. The system input power among three different scenarios is compared in Figure 11. Scenario 1 shows the consistently highest system input power, while Scenario 2 shows the consistently lowest system input power under a similar
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range of airflow rate. Compared with Scenario 1, Scenario 2 saves about 60% system input power at a lower airflow range and 35% of system input power at a higher airflow range. An interesting finding is that the system input power
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in Scenario 3 spreads out between Scenarios 1 and 2, which is different from the simulation results. By comparing measured static pressure, shown in Figure 12(a), Scenario 3 maintained much higher duct static pressure than Scenario 2 along with expected oscillations. This is due to the more severe load diversity among the thermal zones than the one assumed in the simulations. Even though Scenario 2 shows much better energy performance, starved thermal zones occur as indicated in Figure 12(b), where the summation of the total supply airflow rate from all the
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terminal boxes is around 25% lower than the summation of the supply airflow set points. The outdoor air temperature is shown in the same figure using the label on the right side. The results indicate that some thermal zones may suffer thermal problems. On the other hand, the measured airflow rate is well maintained at the airflow
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rate set points by Scenario 3.
Figure 11. Measured system input power comparison among three different scenarios.
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Figure 12. Measured performance comparison among three different scenarios.
CONCLUSION
A comprehensive mathematical model to calculate VFD-motor-fan input power was established by including supply airflow rate, pressure losses, fan efficiency, motor efficiency, and VFD efficiency as inputs in this paper. Simulations were conducted to compare fan operations, system input power, power exponent factors, and system resistance coefficients under three different fan control scenarios: 1) using a constant static pressure set point; 2) resetting the static pressure using total airflow rate (feedback from AHUs); and 3) resetting static pressure using the
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worst zone (feedback from zones). The simulation results show that more than 50% of fan power consumption can be observed in Scenarios 2 and 3 under partial load conditions. In addition, the impact of minimum airflow ratio (MAR) on the fan power energy performance was also investigated. The results show that moderate MAR
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reductions can result in a reasonable amount of power savings. It is also concluded that for the simulation case (the 3hp fan), no significant system input power savings was observed by reducing MAR from 30% to 10% due to the increased efficiency losses from the fan, motor and VFD at low load and speed conditions. Experiments has also approved that up to 60% of system input power savings can be obtained in Scenario 2 (the pressure reset by the total
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airflow rate) compared with Scenario 1. However, the performance of Scenario 3 heavily relies on the distribution of the zone loads. Unbalanced zone loads significantly reduced fan power savings compared with the simulation
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results.
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[9] Wang S, Dynamic simulation of building VAV air-conditioning system and evaluation of EMCS on-line control strategies, Building and Environment 34 (1999): 681-705. [10] ANSI/ASHRAE/IESNA Standard 90.1-2010. Energy Standard for Buildings Except Low-Rise Residential
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Buildings. Atlanta : American Society of Heating, Refrigerating and Air-conditioning Engineers, Inc; 2010 [11] Zhao T, Zhang J, Ma L, Experimental analysis of thermal stability of the pressure control method for a variable flow air-conditioning water system, Building and Environment 70 (2013): 1-9.
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[13] Wang S, Jin X, Model-based optimal control of VAV air-conditioning system using genetic algorithm, Building
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[16] Lawrence Berkeley National Laboratory (LBNL), EnergyPlus Engineering Reference – Air system Fans. 2011;704-724 http://apps1.eere.energy.gov/buildings/energyplus/energyplus_documentation.cfm [17] Branesky B, Song L, Air handling unit level fault signature development using EnergyPlus, Proceedings of
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ASME2012 International Mechanical Engineering Congress and Exposition, IMECE2010-87574, Houston TX. [18] ASHRAE. HVAC Systems and Equipment. Chapter 20.4. Fans. .
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Refrigerating and Air-conditioning Engineers, Inc; 2008, p.20 [19] Wang G, Liu M, Claridge DE,. Development of an Energy Meter Using a Pump Flow Station. ASHRAE Transactions 2010;116(2)
[20] Wang G, Song L, Park S, Estimation of Induction Motor Efficiency under Variable Frequencies.
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A comprehensive mathematical model is built to calculate VFD-motor-fan power. 3 scenarios include no reset, reset by total flow and rest by highest zone demand. The model can generate input curves to EnergyPlus to simulate a VFD-motor-fan system. The simulation results show that more than 50% savings. Experiments show the zone feedback control relies on zone load diversity.
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• • • • •