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Control and energy simulation of variable refrigerant flow air conditioning system combined with outdoor air processing unit
Accepted Manuscript Control and energy simulation of variable refrigerant flow air conditioning system combined with outdoor air processing unit Yonghua Zhu, Xinqiao Jin, Zhimin Du, Xing Fang, Bo Fan PII:
S1359-4311(13)00973-3
DOI:
10.1016/j.applthermaleng.2013.12.076
Reference:
ATE 5294
To appear in:
Applied Thermal Engineering
Received Date: 1 September 2013 Accepted Date: 28 December 2013
Please cite this article as: Y. Zhu, X. Jin, Z. Du, X. Fang, B. Fan, Control and energy simulation of variable refrigerant flow air conditioning system combined with outdoor air processing unit, Applied Thermal Engineering (2014), doi: 10.1016/j.applthermaleng.2013.12.076. 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.
ACCEPTED MANUSCRIPT [Title Page]
Control and energy simulation of variable refrigerant flow air
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conditioning system combined with outdoor air processing unit Yonghua ZHU, Xinqiao JIN*, Zhimin DU, Xing FANG, Bo FAN
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School of Mechanical Engineering, Shanghai Jiao Tong University, Shanghai 200240, P. R. China
author: Xinqiao JIN
*Corresponding
email: [email protected]
*Corresponding
Tel/fax: +86-21-34206774
*Corresponding
mailing address:
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*Corresponding
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Institute of refrigeration and cryogenics, Shanghai Jiao Tong University, No. 800, Rd. Dong-Chuan, Dist. Min-Hang, Shanghai, China.
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Zip/Postal Code: 200240
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Control and energy simulation of variable refrigerant flow air conditioning system combined with outdoor air processing unit Yonghua ZHU, Xinqiao JIN*, Zhimin DU, Xing FANG, Bo FAN
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School of Mechanical Engineering, Shanghai Jiao Tong University, Shanghai 200240, P. R. China
Abstract: A variable refrigerant flow (VRF) unit and outdoor air (OA) processing unit combined air conditioning system is proposed as a solution for ventilation problems in VRF systems. System
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structure and control strategies are addressed. Simulation platform is established based on
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component and sub-system models developed. Control and energy performances analysis are then put forward under various conditions. It is found that the combined system could maintain all the zones at their specific set-points within small errors no matter set-points of the zones are the same or different. In addition, indoor air quality can be ensured. Energy efficiency characteristics of the
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combined system are greatly affected by the OA supply temperature of the OA processing unit, which opens up the opportunity of minimizing the energy consumption of the combined system through optimal control strategy. Results reveal that the best OA supply temperature can be
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obtained through optimizing part load ratio of the OA processing unit to a range in which the
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system operates with high efficiency. Key words: combined air conditioning system; variable refrigerant flow; outdoor air processing; indoor air quality; part load ratio
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Heating, ventilating and air-conditioning
Q
Cooling, W
VRF
Variable refrigerant flow
D
Humidity load, kg/s
VAV
Variable air volume
w
Humidity ratio, kg/kg (dry air)
IDU
Indoor unit
cp
Specific heat capacity, kJ/kg. K
IAQ
Indoor air quality
C
CO2 concentration, ppm
DX
Direct expansion
G
CO2 emission rate of people,
OA
Outdoor air
Occp
Number of occupants
EEV
Electronic expansion valve
PLR
Part load ratio
Subscripts
COP
Coefficient of performance
VRF
M
Mass, kg
OA
outdoor air
m
Mass flow rate, kg/s
com
compressor
V
Volume, m3
fan
supply fan
v
Volumetric flow rate, m /s
i
ith air conditioned zone
T
Temperature, ℃
s
supply air
W
Input power, W
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1. Introduction
variable refrigerant flow
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HVAC
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Nomenclature
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It is estimated by recent simulation studies and field surveys that heating, ventilating and air-conditioning (HVAC) systems consume as much as 40% of the total electricity use in the office buildings [1], which arouses worldwide attention to study the methods for reducing energy use of the air conditioning systems. Conventional HVAC systems, such as all-air variable air volume (VAV) systems, have been found several deficiencies along with the wide application in many types of buildings. These deficiencies [2, 3] including occupying much space, inflexibility in maintenance and commissioning, energy wasting for offsetting the extra cooling by reheating, etc. The variable refrigerant flow (VRF) system, first introduced to the market in Japan, now has 3
ACCEPTED MANUSCRIPT earned a wide application in both residential and commercial buildings, due mostly to high energy efficiency under part load condition and individualized thermal comfort capability [4-7]. In addition, duct losses in the VRF systems can be almost ignored due to the in-space location of the
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indoor units [3]. Up to date, there have been many literatures about experimental and numerical studying of the VRF systems. Xia et al. [8] developed a test rig for a three-pipe VRF system having five
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indoor units. It was found that the COP of the system did not vary too much according to the part
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load ratio. Aynur et al. [9, 10] evaluated the performance of VRF systems by experimental and simulation approaches. Choi and Kim [11] investigated the capacity modulation method by varying the indoor loads, the electronic expansion valve (EEV) opening and the compressor speed. Shi et al. [12] and Shao et al. [13] developed a fluid network model to simulate the performance of
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the three-pipe VRF system. Zhou et al. [4] developed a simulation module for the VRF system based on EnergyPlus by curve-fitting method for predicting and evaluating energy performances of the VRF systems and hence for making comparisons with other types of HVAC systems.
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Similar research can be found in [5] for a water-cooled VRF system. Zhu et al [14, 15] developed
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generic simulation models for the VRF system both in cooling and heating modes, where the generality means the model is component-number independent. Control performance of the VRF systems is another research focus. Masuda et al. [16] developed a control method for a VRF system with two indoor units. Chen et al. [17] investigated the control performance of a triple-evaporator system in which a relative simple model by black-box method was developed first. The simulation results show good control performance in the evaporator temperature and superheat. Shah et al. [18] used the mean void fraction and moving boundary approach to develop 4
ACCEPTED MANUSCRIPT the VRF system model for designing advanced closed-loop controllers. Lin and Yeh [19] addressed the feedback control design for a triple-evaporator air conditioner through systematic identification. The identification procedure produces a low-order, linear model suitable for the
controlled simultaneously to make the system operate in high efficiency.
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controller design. The studies suggested that the EEV and the compressor speed should be
Besides the research endeavors that have been made for field testing, energy and control
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simulation of the VRF system, combining VRF systems with other systems are also studied
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aiming at taking advantages of different systems. Aynur et al. [20, 21] studied the performances of integration of VRF and heat pump desiccant both in cooling and heating modes. The integration was found to be promising in terms of energy saving and better indoor thermal comfort. Karunakaran and Parameshwaran [22, 23] developed a simulation model and a test rig for joint
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control of the VRF system and ventilation devices. The system actually is a variant of VAV system that the air is processed by two evaporators (in cooling mode) of the VRF laid in series. Jiang et al [24] proposed a solid desiccant heat pump and VRF combined air conditioning system to form a
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temperature and humidity independent control system.
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Literatures [3, 20, 21] also pointed out that the shortcoming of the VRF system lacking ventilation ability has not been solved thoroughly. The VRF system can hardly ensure the indoor air quality (IAQ) alone without additional ventilating devices. In most circumstances, an energy recovery ventilator [25, 26] will be installed accompanying with the VRF system for inducing efficient amount of ventilation air while recovering energy from the exhaust air stream in order to reduce the OA load. However, the energy recovery ventilator usually is not controllable for the outdoor airflow [27], thus fixed rather than suitable outdoor airflow is supplied to the air 5
ACCEPTED MANUSCRIPT conditioning zones, and it often results in energy wasting, which will be worse in the case of applications with varying occupancy. Inadequate combination of VRF with the ventilation system (such as the energy recovery ventilator) either results in poor indoor thermal comfort, or more
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energy consumption due to the additional OA load [28]. That’s why the combination of the VRF system and the ventilation system gains importance in practical applications. Unfortunately,
researchers in the VRF field focus their interests mostly on features of the equipment itself [6-9,
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17-19, 29, 30], e.g. simulation or experiment research of the variable speed compressor and the
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electric expansion valve (EEV) and novel control logic, etc. Still very few attentions [15, 25, 26] are paid to the combination of VRF system with ventilation system. With above considerations, a new air conditioning system combining VRF unit and OA processing unit is proposed aiming to take advantages of both parts. The OA processing unit cools and dehumidifies OA through an
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inherent direct expansion (DX) unit, and supplies the processed OA with low humidity to the indoor as the supply air in cooling mode. And the parallel VRF system accommodates the remaining loads, including the indoor loads and possible OA load.
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In current research, the combined air conditioning system (noted as combined system
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hereinafter) is intended to be investigated in terms of control strategy, energy performance, thermal comfort and IAQ in cooling mode. Based on the system and sub-system models of VRF unit and other related components developed and validated in previous researches [14, 15, 31, 32], a TRNSYS [33] based simulation platform of the combined air conditioning system is developed first. Simulation results are displayed and discussed following. The remainder of the paper is organized as follows: Section 2 describes simulation methodology of the combined system after the structure and control strategies are illustrated. 6
ACCEPTED MANUSCRIPT Section 3 displays results and discussions of the combined system under the control strategies. Energy and efficiency characteristics of the combined system are discussed in Section 4. And Section 5 concludes this study.
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2. Simulation of the combined system and its controls 2.1 Description of the combined system and control designing
A schematic diagram of the proposed combined system is shown in Fig. 1. The system can be
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divided into two parts structurally, VRF part and OA processing part. The VRF part consists of
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one outdoor unit and multiple indoor units (IDU), varying the refrigerant flow rate with the help of the variable speed compressor in the outdoor unit and the electronic expansion valve (EEV) located in each IDU to match the cooling/heating load in order to maintain the zone air temperature at the indoor set-point. The OA processing part consists of a DX unit and VAV boxes.
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The OA conditioned by DX unit is supplied into zones by the supply fan through the VAV boxes. The VAV box regulates the conditioned OA flow corresponding to CO2 contaminant concentration for maintaining IAQ, since many studies [1, 31] have found a worsening of IAQ outcomes at
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higher CO2 concentrations. In addition, there is an enthalpy wheel between the exhausted air and
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the OA to recover energy.
The control strategies of the combined system are also shown in Fig. 1. The control strategies are designed for three targets: a) maintaining thermal comfort of the multi-zone, b) maintaining acceptable indoor air quality, and c) ensuring reliable operating of the VRF unit and the OA processing unit. For VRF part, two control loops are set for accommodating the varying loads. The zone temperature control loop samples the return air temperature of each zone to regulate the opening 7
ACCEPTED MANUSCRIPT of EEV in the IDU. The VRF capacity control loop samples superheated degree of the overall suction vapor to regulate the speed of the compressor. Mass flow rate of the refrigerant supplied to the evaporators are controlled by the two loops.
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For OA processing part, four control loops are set for processing and regulating sufficient OA for ventilation. OA flow rate control is accomplished by combination of VAV box control and
supply fan control based on concept of demanded control ventilation (DCV) [31]. When dampers
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of the zones are regulated, supply air static pressure in the air duct may change, and the supply fan
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control loop modulates the fan speed to maintain constant supply static pressure. In addition, economizer cycle technique [34] is also incorporated and it works when the OA enthalpy is smaller than EA enthalpy in cooling mode. The other two control loops are designed for conditioning OA. The OA supply temperature is maintained by the EEV opening control loop
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through regulating the opening of the EEV. The DX capacity control loop samples the superheat degree to adjust the compressor speed.
2.2 Simulation of the combined system
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The simulation of the proposed combined system is setup based on authors’ previous works
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[14, 15, 31, 32]. The developed simulator integrates the individual sub-system or component models into a complete system by use of TRNSYS software. It should be noted here that the ready-made multi-zone building model provided by TRNSYS, i.e. Type56, is used for simulating thermal performances of each zone. However, parameters and settings of the building, including location, materials, dimensions, etc., are provided accordingly, which will be detailed discussed and modeled in the following sub-section. Both the VRF unit and the DX unit are air cooled refrigeration systems. In addition, the DX 8
ACCEPTED MANUSCRIPT unit can be treated as a VRF unit with only one evaporator. It is reasonable to develop generic models (i.e. independent of component numbers) for the VRF unit. In authors’ previous work [14], several simulation models for the VRF unit were developed and validated. All the models were
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proved of fast computing and evaporator-number independence, and showed good ability for control analysis. One of the VRF models is used in the simulation of the combined system.
2.3 Building description and air conditioned zone model
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An office building is selected to accommodate the combined system. The building has six
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floors in total above ground. Each floor of the building is divided into six conditioned thermal zones, corresponding to four outdoor exposures (east, west, south and north), an interior zone (including a ring-shaped corridor) and a center core, as shown in Fig. 2. The center core is space for elevators and stairs. The air conditioned zones are numbered from zone 1 to zone 25 for
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conveniently identifying their locations. A summary of the key parameters of the building is listed in Table 1.
By considering each zone as a control volume, ordinary differential equations [31, 32], which
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express the transient behaviors of temperature, humidity and CO2 concentration, are derived and
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described as follows. It is noted here that balances of the sensible load and the latent load are separately considered for having an insight into their variations and in accordance with the modeling methodology of TRNSYS.
M ic p
Mi
Vi
dTi = ms ,VRF ,i c p (Ts ,VRF ,i − Ti ) + ms ,OA,i c p (Ts ,OA − Ti ) + Qi dτ
dwi = ms ,VRF ,i ( ws ,VRF ,i − wi ) + ms ,OA,i ( ws ,OA − wi ) + Di dτ
dCi = vs ,i ( Cs ,i − Ci ) + G ⋅ Occpi dτ
(1)
(2)
(3) 9
ACCEPTED MANUSCRIPT Where, M is the mass of the air, V is the volume of the zone, T is the indoor temperature, ms is the mass flow rate of supply air, Ts is the temperature of supply air, Q is the internal load, w is the indoor humidity ratio, ws is the humidity ratio of supply air, D is the internal humidity load, C is
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the indoor CO2 concentration, Cs is the CO2 concentration of supply air (360 ppm is used), vs is the volume flow rate of supply air, Occp is the number of occupants and G is the amount of CO2
emission rate of people (5×10-6 m3/s is used). The subscript VRF, OA and i represent VRF part,
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OA processing part and ith zone, respectively.
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The internal load (Q) and the internal humidity load (D), generated by occupant, lighting and equipment, are simulated by offering operation schedules to mimic the practical variations. Other factors, including solar radiation, heat transfer through window and envelops are not manually interacted. Table 2 describes the methodology of heat gain calculation for occupant, lighting and
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equipment. LightMax, EqMax and OccMax denote designed lighting load per square, designed equipment load and designed occupancy number, respectively. LSchedule, ESchedule and OccSchedule denote possibility of maximum lighting, number of operating equipment and
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occupancy density, respectively. OHeat denotes heat generation rate per person.
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3. Control performances of the combined system 3.1 Test conditions
Zones 7 to 12 as shown in Fig. 2 are selected for the study because of good representing of internal and external factors such as orientations, radiations, etc. Heat generated by lighting and equipment are designed the same for each zone. Schedules of lighting and equipment are shown in Fig. 3. The profile of occupancy in each zone is not the same and is simply distinguished by multiplying a different occupant factor onto a base occupant density, which is shown in Fig. 3 as 10
ACCEPTED MANUSCRIPT well. The occupant factors for zone 7 to 12 are 0.95, 1.0, 0.8, 1.05, 1.10 and 0.90, respectively. The light schedule (0~1) describes the possibility of maximum lighting to mimic those offices with intensive light that one can turn (some of) them on and off conveniently. The corridor
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adjacent to these zones is assumed to have an independent air conditioner, which has a thermostat set to 24℃ in cooling mode. It is assumed that there is no air exchange between different floors in this paper. 30th July of Shanghai, whose TMY (Typical Meteorological Year) weather profile is
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supposed to be operated during 8 a.m. to 20 p.m.
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shown in Fig. 4, is chosen for representing the typical cooling conditions. The combined system is
3.2 Control performances
Two cases, the same and different temperature set-points of the zones are investigated. For the case of the same set-point, the set-points of all the 6 zones are 24℃, and the OA supply
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temperature set-point is also 24℃. For the case of different set-point, the OA supply temperature set-point is 20℃, zone 7 and zone 12 are 24℃, zones 8 is 25℃, zone 9 and zone 10 are 26℃. In addition, in the case of different set-point, zone 11 is not controlled that the IDU is stopped to test
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how well the combined system adapts to different amount of operating indoor units.
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3.2.1 Control performance of VRF part Figs. 5 and 6 show the control performances of the VRF part in the case of the same set-point and different set-point, respectively. As shown in Fig. 5a and 6a, all the zones can be maintained at their specific set-points, and the control is also stable no matter the set-points change or not. There are tiny peaks and valleys in the temperature profiles of the zones in either of the cases, due mostly to the continuously varying OA flow rate. For zone 11 in the case of different set-point, the temperature changes freely and 11
ACCEPTED MANUSCRIPT rationally due to load variation since it is not controlled. The highest temperature achieves about 30℃ at early afternoon when the cooling load is comparatively large during the day. The zone relative humidity is not directly controlled. Instead, in the process of meeting the sensible load of
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each zone, dew-point temperature of the air is forced below the tube wall temperature, coincidentally providing dehumidification as a result. However, as shown in Fig. 5b and 6b, relative humidity of all the controlled air conditioned zones can meet the thermal comfort
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requirements as they are all maintained at about 40% ~ 65%. The separating of the relative
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humidity in the case of the same set-point is caused by the different occupant condition in each zone. The high temperature and relative humidity of zone 11 in the case of different set-point imply the importance of air conditioning for buildings in summer in Shanghai. Fig. 5c and 5d depict the changing behaviors of the EEVs’ opening and the compressor speed
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(nominated by dividing the rated value, 3000rpm) in the case of the same set-point, respectively. The corresponding variations in the case of different set-point are depicted in Fig. 6c and 6d, respectively. In both cases, all the EEVs except that in zone 11 in the case of different set-point
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quickly respond to the required cooling requirement, which is represented by the difference
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between the zone air return temperatures and their set-points and the change rates of the zone temperatures. In addition, the compressor speed also changes effectively in response to changes in superheated degree caused by variation of EEVs’ opening so that the compressor capacity can closely match the total loads. 3.2.2 Control performance of OA processing part The control loops of the OA processing part are designed for ensuring sufficient ventilation air in order to reduce CO2 concentration, or in other words, to maintain the IAQ. Since CO2 12
ACCEPTED MANUSCRIPT generation is solely related to the number of occupants that it is independent of zone temperature and humidity, OA flow rate as well as CO2 concentration in either of the two cases are the same. Therefore, the control performances in OA processing part will be illustrated here only based upon
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the case of different set-point (Figs. 7 and 8). As is shown in Fig. 7a, the OA flow rate of the combined system varies gradually with the
time. It is adjusted according to the variation of the occupants. The rising up and descending down
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of the OA flow rate and the occupancy are almost synchronized as a result. The valley in early
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afternoon reflects rational absence of the occupants for lunch, rest, etc. The least minimum OA mass flow rate in this research is set for ensuring basic sanitary condition of the floor when occupants are absent.
The OA flow rate is a result of the combined activities of the VAV box control and the supply
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fan control. Fig. 7b and 7c depict the variations of the air damper in each VAV box and the supply fan frequency. The air dampers are all regulated in rational positions to allow appropriate amount of OA to maintain the zone CO2 concentration at or below the limit, 1000 ppm. The supply fan
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frequency is increased or decreased with the openings of the air dampers to maintain the static
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pressure at the set-point, as shown in Fig. 7c. On the other hand, during most of operation time, the openings of the dampers are all smaller than 75%, it means that there is potential for energy-saving optimal control, i.e., the supply fan frequency could be optimized (reduced) to save energy.
The indoor CO2 concentration as shown in Fig. 7d is well controlled at the IAQ-limit during most of the operation time. The reason of the exception of a short time at the beginning can be explained by the hysteresis of the control system. The tiny peaks and valleys of the CO2 13
ACCEPTED MANUSCRIPT concentration profiles are almost synchronized with the OA flow rate, which implies that the more stable the occupancy variation is, the better control of the OA flow rate as well as the CO2 concentration will be.
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Fig. 8 shows the control performances of the DX unit, including the OA supply temperature, opening of the EEV and speed of the compressor. Fig. 8a shows that the OA supply temperature
opening and compressor speed, as shown in Fig. 8b.
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can be quickly and well controlled at the set-point, with the help of quickly adjusting of the EEV
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In conclusion, test results show that the combined system has good ability and performances responding to various conditions. The control strategies are effective and efficiently implemented. In addition, the demand of individual control could also be met by applying this combined system.
4. Energy efficiency characteristics of the combined system
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4.1 Energy consumption under various conditions
Although the OA processing part is designed in basic purpose to supply and process sufficient OA for maintaining IAQ, it has ability to accommodate part of the indoor loads when
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the OA supply temperature is lower enough. In that way, there will be a trade-off between the
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loads taken by the VRF part and by the OA processing part. However, energy consumption of the combined system will not be the same due to varying efficiencies of the individual parts under different load conditions. From the energy saving point of view, the operation strategy that can provide sufficient cooling but consume the least energy will be preferred. Clearly, energy consumption decreases with increasing of COP. Therefore, the energy consumption of the combined system can be reduced or even minimized when both the VRF unit and the OA processing unit operates in the most efficient way. However, it is not easy to make both the VRF 14
ACCEPTED MANUSCRIPT unit and the OA processing unit operate with highest COP simultaneously before the characteristics of energy efficiency are completely understood. To understand the characteristics of energy efficiency of the combined system under different
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conditions, a group of tests (GP-A) are carried out first. Different OA supply temperature set-point, from 18 to 26℃ with a 2℃ increment, is selected for each test, with all the zones’ set-point 24℃. Other testing conditions, including load settings, zone locations, and weather profile, are not
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changed.
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Figs. 9 and 10 show test results of the cooling capacity and energy consumption of the VRF unit and the OA processing unit in GP-A tests, where the cooling capacity and the energy consumption are calculated by integral of the cooling and input power over the operating time, respectively.
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As shown in Fig. 9, the total cooling capacity of the combined system is the same in all the test cases. The cooling capacity of the OA processing part decreases with the increasing set-point of the OA supply temperature, while the cooling capacity of the VRF part acts in a reverse way.
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As shown in Fig. 10, the energy consumption of the VRF unit increases with increasing
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set-point of OA supply temperature for taking more OA loads. However, the energy consumption of the OA processing unit is not monotonely changing with the OA supply temperature. It decreases to the lowest value and then increases with increasing set-point of the OA supply temperature. It means that there is minimum energy consumption for the combined system, i.e., there is a best OA supply temperature that can minimize the energy consumption of the combined system. On the other hand, another group of tests (GP-B) are further carried out with the OA supply 15
ACCEPTED MANUSCRIPT temperature set-point fixed (20℃) and the rated capacity of the OA processing unit (DX unit) varying from 50% to 150% of that used in GP-A in a 25% increment. Other conditions keep the same as GP-A. The results are shown in Fig. 11.
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As shown in Fig. 11, the energy consumption of the OA processing unit decreases to the lowest value when the current DX capacity (i.e. the 100% case) is used. It then increases with increasing DX capacity. It means that the DX capacity affects the energy consumption of the
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combined system. We should carefully select DX capacity for practical applications. Actually, the
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current DX capacity has been determined in GP-A tests for high energy efficiency by trial and error method. However, when the DX capacity is determined, the energy consumption of the combined system is mainly affected by the OA supply temperature, as illustrated in GP-A tests. 4.2 Relationship between COP and PLR
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The interesting results of the total energy consumption in both GP-A and GP-B can be further explained by the relationships between COP and PLR, where they are calculated as follows.
Q Wcom + W fan
PLR =
Q Qrated
(4)
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COP =
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(5)
Where Q and Qrate denote actual and rated cooling of the VRF unit, respectively. W denotes input power, and subscript com and fan denote compressor and supply fan, respectively. It should be noted here that Q and W used here are instantaneous values. Fig. 12 and 13 show the relationships between COP and PLR of the VRF unit and the OA processing unit in GP-A, respectively, where the PLR is calculated as the ratio of the provided cooling to the rated capacity of the individual unit. Generally, for both units, COP decreases with 16
ACCEPTED MANUSCRIPT increment of PLR, and vice versa. When PLR locates in the range from 0.4 to 0.7, COP attains its maximum, which is in accordance with the conclusions of experimental studies [35]. COP falls to the lowest when PLR gets its value larger than 0.8. However, in cases of high OA supply
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temperature set-point (e.g. 26℃), COP of OA processing unit increases or decreases in accordance with PLR strictly, which is caused by the DX compressor having to operate at its minimum value due to relatively small OA load. PLR and COP of the VRF unit change in a narrower range
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comparing to that of the OA processing unit. In other words, OA supply temperature has
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significantly influences on operation of the OA processing unit, and the total energy consumption of the combined system is greatly impacted as a result.
The relationship between COP and PLR of the OA processing unit in GP-B is shown in Fig. 14. That of the VRF unit is not shown again since the operation of the VRF unit is not affected
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when the OA supply condition is fixed. As shown in Fig. 14, COP of the OA processing unit is improved as the DX capacity is decreased during early morning and late afternoon periods when the loads are small. However, the COP is not improved with the decreasing of DX capacity at
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other times. On the other hand, the COP decreases significantly as the DX capacity is increased,
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especially when the capacity is 1.5 times larger. And it is the reason that the energy consumption is increased dramatically in these cases. Specifically, it can be seen from Fig. 13b and 14b that, in most of the operation time, PLR of the OA processing unit locates in the range from 0.4 to 0.7 when the OA supply temperature set-point is 20. So it results in high efficiency operation of the OA processing unit among all the cases. From this point of view, energy consumption of the combined system can be reduced by suitable OA supply temperature set-point, i.e., OA supply temperature set-point could be 17
ACCEPTED MANUSCRIPT optimized according to the operation conditions. In conclusion, results show that the energy efficiency of the VRF unit as well as the OA processing unit changes with variation of load. It reaches the highest when PLR locates in certain
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range, which indicates that further optimization can be realized by making both the VRF unit and the OA processing unit operate in the highest efficiency range.
5. Conclusions and future work
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Simulations of a new air conditioning system combining VRF with OA processing unit in
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cooling conditions are presented. The illustration of the combined system has been introduced and its control loops has been designed. The simulation methodology is developed and the tests are carried out to validate the control performances and energy efficiency characteristics of the combined system.
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The simulation tests results show that all the zones of the combined system could be maintained at their specific set-points within a small error after the control is stable no matter the set-points are the same or different, and no matter the amount of operating indoor units changes or
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not. Besides, the IAQ can be ensured.
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The relationship between COP and PLR indicates that further optimization can be realized by making both the VRF unit and the OA processing unit operate in the high efficiency range. The results also show that there is a best OA supply temperature set-point that can optimize the energy consumption of the combined system. As the OA processing unit affects the operation of the combined system more significantly than the VRF unit, the optimization can be realized by simply controlling the PLR of the OA processing unit to the most efficient system-operation range. In this paper, characteristics of the combined system under regular control strategies are 18
ACCEPTED MANUSCRIPT simulated and analyzed. Preliminary study about optimal control of the combined system is also carried out. Future work could be put forward in several aspects, especially in online optimization
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of the OA supply temperature. The results will be presented in subsequent papers.
Acknowledgements
The research work presented in this paper is financially supported by National Natural Science
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Foundation of China (No. 51376125) and National Natural Science Foundation of China (No.
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51376126).
References
[1] K.J. Chua, S.K. Chou, W.M. Yang, J. Yan, Achieving better energy-efficient air conditioning – A review of technologies and strategies, Appl. Energy 104 (2013) 87-104.
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[2] J.W. Jeong, S.A. Mumma, W.P. Bahnfleth, Energy Conservation Benefits of a Dedicated OA System with Parallel Sensible Cooling by Ceiling Radiant Panels, ASHRAE Trans. 109 (2003) 627-636.
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[3] W. Goetzler, Variable refrigerant flow systems, ASHRAE J. 49 (2007) 24-31.
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[4] Y.P. Zhou, J.Y. Wu, R.Z. Wang, S. Shiochi, Energy simulation in the variable refrigerant flow air-conditioning system under cooling conditions, Energ. Build. 39 (2007) 212-220. [5] Y.M. Li, J.Y. Wu, S. Shiochi, Modeling and energy simulation of the variable refrigerant flow air conditioning system with water-cooled condenser under cooling conditions, Energ. Build. 41 (2009) 949-957. [6] X.B. Liu, T.Z. Hong, Comparison of energy efficiency between variable refrigerant flow systems and ground source heat pump systems, Energ. Build. 42 (2010) 584-589.
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ACCEPTED MANUSCRIPT [7] T.N. Aynur, Y. Hwang, R. Radermacher, Simulation comparison of VAV and VRF air conditioning systems in an existing building for the cooling season, Energ. Build. 41 (2009) 1143-1150.
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[8] J.J. Xia, E. Winandy, B. Georges, J. Lebrun, Experimental analysis of the performances of variable refrigerant flow systems, Build. Serv. Eng. Res. T. 25 (2004) 17-23.
[9] T.N. Aynur, Y. Hwang, R. Radermacher, Experimental evaluation of the ventilation effect on
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the performance of a VRV system in cooling mode - Part I, experimental evaluation,
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HVAC&R Res. 14 (2008) 615-630.
[10] T.N. Aynur, Y. Hwang, R. Radermacher, Simulation evaluation of the ventilation effect on the performance of a VRV system in cooling mode - Part II, simulation evaluation. HVAC&R Res. 14 (2008) 783-795.
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[11] J.M. Choi, Y.C. Kim, Capacity modulation of an inverter-driven multi-air conditioner using electronic expansion valves, Energy 28 (2003) 141-155. [12] W.X. Shi, S.Q. Shao, X.T. Li, X.F. Peng, X.D. Yang, A network model to simulate performance of variable refrigerant volume refrigerant systems, ASHRAE Tran. 109 (2003) 61–68.
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[13] S.Q. Shao, W.X. Shi, X.T. Li, Q.S. Ye, Simulation model for complex refrigeration systems based on two-phase fluid network-Part I: Model development, Int. J. Refrig. 31 (2008)
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490-499.
[14] Y.H. Zhu, X.Q. Jin, Z.M. Du, B. Fan, S.J. Fu, Generic simulation model of multi-evaporator variable refrigerant flow air conditioning system for control analysis, Int. J. Refrig. 36 (2013) 1602-1615. [15] Y.H. Zhu, X.Q. Jin, Z. M. Du, B. Fan, X. Fang, Simulation of variable refrigerant flow air conditioning system in heating mode combined with outdoor air processing unit, Energ. Build. 68 (2014) 571-579. [16] M. Masuda, K. Wakahara, K. Matsui, Development of a multi-split system air conditioner for 20
ACCEPTED MANUSCRIPT residential use, ASHRAE Trans. 97 (1991) 127-131. [17] W. Chen, X.X. Zhou, S.M. Deng, Development of control method and dynamic model for multi-evaporator air conditioners (MEAC), Energ. Convers. Manag. 46 (2005) 451-465.
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[18] R. Shah, A.C. Alleyne, C.W. Bullard, Dynamic modeling and control of multi-evaporator air conditioning systems, ASHRAE Trans. 110 (2004) 109-119. [19] J.L. Lin, T.J. Yeh, Identification and control of multi-evaporator air conditioning systems, Int. J. Refrigeration 30 (2007) 1374-1385. [20] T.N. Aynur, Y. Hwang, R. Radermacher, Integration of variable refrigerant flow and heat
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pump desiccant systems for the heating season, Energ. Build. 42 (2010) 468-476.
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[21] T.N. Aynur, Y. Hwang, R. Radermacher, Integration of variable refrigerant flow and heat pump desiccant systems for the cooling season, Appl. Therm. Eng. 30 (2010) 917-927. [22] R. Parameshwaran, R. Karunakaran, C.V.R. Kumar, S. Iniyan, Energy conservative building air conditioning system controlled and optimized using fuzzy-genetic algorithm, Energ. Build.
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42 (2010) 745-762.
[23] R. Karunakaran, S. Iniyan, G. Ranko, Energy efficient fuzzy based combined variable refrigerant volume and variable air volume air conditioning system for buildings, Appl. Energ.
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87 (2010) 1158-1175.
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[24] Y. Jiang, T.S. Ge, R.Z. Wang, Performance simulation of a joint solid desiccant heat pump andvariable refrigerant flow air conditioning system in EnergyPlus, Energ. Build. 65 (2013) 220-230.
[25] Y.M. Li, J.Y. Wu, Energy simulation and analysis of the heat recovery variable refrigerant flow system in winter, Energ. Build. 42 (2010) 1093-1099. [26] W. Hunt, Variable Refrigerant Flow-Heat Recovery Performance Characterization, in: 2012 ACEEE Summer Study on Energy Efficiency in Buildings, August 12-17, Pacific Grove, 21
ACCEPTED MANUSCRIPT California, USA, pp. 138-147. [27] O.O. Abe, C.J. Simonson, R.W. Besant, W. Shang, Effectiveness of energy wheels from transient measurements. Part I: Prediction of effectiveness and uncertainty, Int. J. Heat Mass
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Transf. 49 (2006) 52-62. [28] T.N. Aynur, Variable refrigerant flow systems: A review, Energ. Build. 42 (2010) 1106-1112. [29] Y.P. Zhou, J.Y. Wu, R.Z. Wang, S. Shiochi, Simulation and experimental validation of the
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variable-refrigerant-volume (VRV) air-conditioning system in EnergyPlus, Energ. Build. 40
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(2008) 1041-1047.
[30] B. Shen, C.K. Rice, Multiple-zone variable refrigerant flow system modeling and equipment performance mapping, ASHRAE Trans. 118 (2012) 420-427.
[31] X.Q. Jin, H.G. Ren, X.K. Xiao, Prediction-based online optimal control of OA of multi-zone
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VAV air conditioning systems, Energ. Build. 37 (2005) 939-944.
[32] S.W. Wang, X.Q. Jin, Model-based optimal control of VAV air-conditioning system using genetic algorithm, Build. Environ. 35 (2000) 471-487.
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[33] S.A. Klein et al, TRNSYS-A Transient Simulation Program, User Manual version 17.0, 2009.
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[34] Y. Ye, W. Li, Energy analysis on VAV system with different air-side economizers in China, Energ. Build. 42 (2010) 1220-30. [35] W.H. Xue, P.L. Chen, C.J. Liu, The experimental study of the operating characteristic and energy consumption of the VRV air-conditioning system, Energ. Conserv. Tech. 21 (2003) 3-5. (In Chinese)
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List of Figure captions
Fig. 2 Typical floor plan of the building (unit in drawing: meter) Fig. 3 Base occupant density and schedules of equipment and lighting
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Fig. 4 TMY weather profile of Shanghai in 30th July
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Fig. 1 Schematic and control designing of the combined system
Fig. 5 Control performances of the VRF part in the same set-point case: (a) temperature, (b)
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relative humidity, (c) EEV opening, (d) compressor speed
Fig. 6 Control performances of the VRF part in different set-point case: (a) temperature, (b) relative humidity, (c) EEV opening, (d) compressor speed
Fig. 7 Control performances of OA flow: (a) OA flow rate, (b) air damper opening, (c) supply fan
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Fig. 8 Control performances of DX unit: (a) OA supply temperature, (b) Compressor speed and
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Fig. 9 Cooling capacity of the combined system in GP-A tests
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Fig. 10 Energy consumption of the combined system in GP-A tests Fig. 11 Energy consumption comparison of the combined system in GP-B tests Fig. 12 PLR and COP of VRF unit in GP-A tests Fig. 13 PLR and COP of OA processing unit in GP-A tests Fig. 14 PLR and COP of OA processing unit in GP-B tests
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Building location Building type and storeys Gross floor area (air conditioned area) Typical floor plan Zone height (m) Solid wood door size Windows and shading
Shanghai, China Office building, 6-story above ground 4700 m2 Area=28 m×28 m, floor-to-floor height=3.5m 2.7 m Width × height =1.0 m×2.0 m Low-e double pane glazing. Window width × height = 1.8 m×1.5 m; sill height=0.80 m; No shading device Aerated concrete wall, thickness=310 mm. Concrete blocks, thickness=200 mm. Reinforced concrete, thickness=120 mm.
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External wall Internal wall Floor
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Description
Lighting
Lightload=LightMax × LSchedule × ZoneArea LightMax =20W/m2; LSchedule (0~1). Equipload=EqMax × ESchedule EqMax =120W; ESchedule (0~6). Occpload= OccMax × OccSchedule×OHeat OccMax=6; Seated, very light activity, OHeat=65W.
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Occupant
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> A VRF and ventilation combined system is proposed to take merits of both parts; > A simulation platform is developed for the combined system; > The combined system could maintain thermal comfort and indoor air quality; > Characteristics of system energy efficiency are revealed for further optimal control;
S1359-4311(13)00973-3
DOI:
10.1016/j.applthermaleng.2013.12.076
Reference:
ATE 5294
To appear in:
Applied Thermal Engineering
Received Date: 1 September 2013 Accepted Date: 28 December 2013
Please cite this article as: Y. Zhu, X. Jin, Z. Du, X. Fang, B. Fan, Control and energy simulation of variable refrigerant flow air conditioning system combined with outdoor air processing unit, Applied Thermal Engineering (2014), doi: 10.1016/j.applthermaleng.2013.12.076. 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.
ACCEPTED MANUSCRIPT [Title Page]
Control and energy simulation of variable refrigerant flow air
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conditioning system combined with outdoor air processing unit Yonghua ZHU, Xinqiao JIN*, Zhimin DU, Xing FANG, Bo FAN
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School of Mechanical Engineering, Shanghai Jiao Tong University, Shanghai 200240, P. R. China
author: Xinqiao JIN
*Corresponding
email: [email protected]
*Corresponding
Tel/fax: +86-21-34206774
*Corresponding
mailing address:
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*Corresponding
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Institute of refrigeration and cryogenics, Shanghai Jiao Tong University, No. 800, Rd. Dong-Chuan, Dist. Min-Hang, Shanghai, China.
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Zip/Postal Code: 200240
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Control and energy simulation of variable refrigerant flow air conditioning system combined with outdoor air processing unit Yonghua ZHU, Xinqiao JIN*, Zhimin DU, Xing FANG, Bo FAN
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School of Mechanical Engineering, Shanghai Jiao Tong University, Shanghai 200240, P. R. China
Abstract: A variable refrigerant flow (VRF) unit and outdoor air (OA) processing unit combined air conditioning system is proposed as a solution for ventilation problems in VRF systems. System
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structure and control strategies are addressed. Simulation platform is established based on
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component and sub-system models developed. Control and energy performances analysis are then put forward under various conditions. It is found that the combined system could maintain all the zones at their specific set-points within small errors no matter set-points of the zones are the same or different. In addition, indoor air quality can be ensured. Energy efficiency characteristics of the
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combined system are greatly affected by the OA supply temperature of the OA processing unit, which opens up the opportunity of minimizing the energy consumption of the combined system through optimal control strategy. Results reveal that the best OA supply temperature can be
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obtained through optimizing part load ratio of the OA processing unit to a range in which the
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system operates with high efficiency. Key words: combined air conditioning system; variable refrigerant flow; outdoor air processing; indoor air quality; part load ratio
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Heating, ventilating and air-conditioning
Q
Cooling, W
VRF
Variable refrigerant flow
D
Humidity load, kg/s
VAV
Variable air volume
w
Humidity ratio, kg/kg (dry air)
IDU
Indoor unit
cp
Specific heat capacity, kJ/kg. K
IAQ
Indoor air quality
C
CO2 concentration, ppm
DX
Direct expansion
G
CO2 emission rate of people,
OA
Outdoor air
Occp
Number of occupants
EEV
Electronic expansion valve
PLR
Part load ratio
Subscripts
COP
Coefficient of performance
VRF
M
Mass, kg
OA
outdoor air
m
Mass flow rate, kg/s
com
compressor
V
Volume, m3
fan
supply fan
v
Volumetric flow rate, m /s
i
ith air conditioned zone
T
Temperature, ℃
s
supply air
W
Input power, W
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1. Introduction
variable refrigerant flow
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3
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HVAC
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Nomenclature
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It is estimated by recent simulation studies and field surveys that heating, ventilating and air-conditioning (HVAC) systems consume as much as 40% of the total electricity use in the office buildings [1], which arouses worldwide attention to study the methods for reducing energy use of the air conditioning systems. Conventional HVAC systems, such as all-air variable air volume (VAV) systems, have been found several deficiencies along with the wide application in many types of buildings. These deficiencies [2, 3] including occupying much space, inflexibility in maintenance and commissioning, energy wasting for offsetting the extra cooling by reheating, etc. The variable refrigerant flow (VRF) system, first introduced to the market in Japan, now has 3
ACCEPTED MANUSCRIPT earned a wide application in both residential and commercial buildings, due mostly to high energy efficiency under part load condition and individualized thermal comfort capability [4-7]. In addition, duct losses in the VRF systems can be almost ignored due to the in-space location of the
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indoor units [3]. Up to date, there have been many literatures about experimental and numerical studying of the VRF systems. Xia et al. [8] developed a test rig for a three-pipe VRF system having five
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indoor units. It was found that the COP of the system did not vary too much according to the part
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load ratio. Aynur et al. [9, 10] evaluated the performance of VRF systems by experimental and simulation approaches. Choi and Kim [11] investigated the capacity modulation method by varying the indoor loads, the electronic expansion valve (EEV) opening and the compressor speed. Shi et al. [12] and Shao et al. [13] developed a fluid network model to simulate the performance of
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the three-pipe VRF system. Zhou et al. [4] developed a simulation module for the VRF system based on EnergyPlus by curve-fitting method for predicting and evaluating energy performances of the VRF systems and hence for making comparisons with other types of HVAC systems.
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Similar research can be found in [5] for a water-cooled VRF system. Zhu et al [14, 15] developed
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generic simulation models for the VRF system both in cooling and heating modes, where the generality means the model is component-number independent. Control performance of the VRF systems is another research focus. Masuda et al. [16] developed a control method for a VRF system with two indoor units. Chen et al. [17] investigated the control performance of a triple-evaporator system in which a relative simple model by black-box method was developed first. The simulation results show good control performance in the evaporator temperature and superheat. Shah et al. [18] used the mean void fraction and moving boundary approach to develop 4
ACCEPTED MANUSCRIPT the VRF system model for designing advanced closed-loop controllers. Lin and Yeh [19] addressed the feedback control design for a triple-evaporator air conditioner through systematic identification. The identification procedure produces a low-order, linear model suitable for the
controlled simultaneously to make the system operate in high efficiency.
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controller design. The studies suggested that the EEV and the compressor speed should be
Besides the research endeavors that have been made for field testing, energy and control
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simulation of the VRF system, combining VRF systems with other systems are also studied
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aiming at taking advantages of different systems. Aynur et al. [20, 21] studied the performances of integration of VRF and heat pump desiccant both in cooling and heating modes. The integration was found to be promising in terms of energy saving and better indoor thermal comfort. Karunakaran and Parameshwaran [22, 23] developed a simulation model and a test rig for joint
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control of the VRF system and ventilation devices. The system actually is a variant of VAV system that the air is processed by two evaporators (in cooling mode) of the VRF laid in series. Jiang et al [24] proposed a solid desiccant heat pump and VRF combined air conditioning system to form a
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temperature and humidity independent control system.
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Literatures [3, 20, 21] also pointed out that the shortcoming of the VRF system lacking ventilation ability has not been solved thoroughly. The VRF system can hardly ensure the indoor air quality (IAQ) alone without additional ventilating devices. In most circumstances, an energy recovery ventilator [25, 26] will be installed accompanying with the VRF system for inducing efficient amount of ventilation air while recovering energy from the exhaust air stream in order to reduce the OA load. However, the energy recovery ventilator usually is not controllable for the outdoor airflow [27], thus fixed rather than suitable outdoor airflow is supplied to the air 5
ACCEPTED MANUSCRIPT conditioning zones, and it often results in energy wasting, which will be worse in the case of applications with varying occupancy. Inadequate combination of VRF with the ventilation system (such as the energy recovery ventilator) either results in poor indoor thermal comfort, or more
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energy consumption due to the additional OA load [28]. That’s why the combination of the VRF system and the ventilation system gains importance in practical applications. Unfortunately,
researchers in the VRF field focus their interests mostly on features of the equipment itself [6-9,
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17-19, 29, 30], e.g. simulation or experiment research of the variable speed compressor and the
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electric expansion valve (EEV) and novel control logic, etc. Still very few attentions [15, 25, 26] are paid to the combination of VRF system with ventilation system. With above considerations, a new air conditioning system combining VRF unit and OA processing unit is proposed aiming to take advantages of both parts. The OA processing unit cools and dehumidifies OA through an
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inherent direct expansion (DX) unit, and supplies the processed OA with low humidity to the indoor as the supply air in cooling mode. And the parallel VRF system accommodates the remaining loads, including the indoor loads and possible OA load.
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In current research, the combined air conditioning system (noted as combined system
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hereinafter) is intended to be investigated in terms of control strategy, energy performance, thermal comfort and IAQ in cooling mode. Based on the system and sub-system models of VRF unit and other related components developed and validated in previous researches [14, 15, 31, 32], a TRNSYS [33] based simulation platform of the combined air conditioning system is developed first. Simulation results are displayed and discussed following. The remainder of the paper is organized as follows: Section 2 describes simulation methodology of the combined system after the structure and control strategies are illustrated. 6
ACCEPTED MANUSCRIPT Section 3 displays results and discussions of the combined system under the control strategies. Energy and efficiency characteristics of the combined system are discussed in Section 4. And Section 5 concludes this study.
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2. Simulation of the combined system and its controls 2.1 Description of the combined system and control designing
A schematic diagram of the proposed combined system is shown in Fig. 1. The system can be
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divided into two parts structurally, VRF part and OA processing part. The VRF part consists of
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one outdoor unit and multiple indoor units (IDU), varying the refrigerant flow rate with the help of the variable speed compressor in the outdoor unit and the electronic expansion valve (EEV) located in each IDU to match the cooling/heating load in order to maintain the zone air temperature at the indoor set-point. The OA processing part consists of a DX unit and VAV boxes.
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The OA conditioned by DX unit is supplied into zones by the supply fan through the VAV boxes. The VAV box regulates the conditioned OA flow corresponding to CO2 contaminant concentration for maintaining IAQ, since many studies [1, 31] have found a worsening of IAQ outcomes at
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higher CO2 concentrations. In addition, there is an enthalpy wheel between the exhausted air and
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the OA to recover energy.
The control strategies of the combined system are also shown in Fig. 1. The control strategies are designed for three targets: a) maintaining thermal comfort of the multi-zone, b) maintaining acceptable indoor air quality, and c) ensuring reliable operating of the VRF unit and the OA processing unit. For VRF part, two control loops are set for accommodating the varying loads. The zone temperature control loop samples the return air temperature of each zone to regulate the opening 7
ACCEPTED MANUSCRIPT of EEV in the IDU. The VRF capacity control loop samples superheated degree of the overall suction vapor to regulate the speed of the compressor. Mass flow rate of the refrigerant supplied to the evaporators are controlled by the two loops.
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For OA processing part, four control loops are set for processing and regulating sufficient OA for ventilation. OA flow rate control is accomplished by combination of VAV box control and
supply fan control based on concept of demanded control ventilation (DCV) [31]. When dampers
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of the zones are regulated, supply air static pressure in the air duct may change, and the supply fan
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control loop modulates the fan speed to maintain constant supply static pressure. In addition, economizer cycle technique [34] is also incorporated and it works when the OA enthalpy is smaller than EA enthalpy in cooling mode. The other two control loops are designed for conditioning OA. The OA supply temperature is maintained by the EEV opening control loop
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through regulating the opening of the EEV. The DX capacity control loop samples the superheat degree to adjust the compressor speed.
2.2 Simulation of the combined system
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The simulation of the proposed combined system is setup based on authors’ previous works
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[14, 15, 31, 32]. The developed simulator integrates the individual sub-system or component models into a complete system by use of TRNSYS software. It should be noted here that the ready-made multi-zone building model provided by TRNSYS, i.e. Type56, is used for simulating thermal performances of each zone. However, parameters and settings of the building, including location, materials, dimensions, etc., are provided accordingly, which will be detailed discussed and modeled in the following sub-section. Both the VRF unit and the DX unit are air cooled refrigeration systems. In addition, the DX 8
ACCEPTED MANUSCRIPT unit can be treated as a VRF unit with only one evaporator. It is reasonable to develop generic models (i.e. independent of component numbers) for the VRF unit. In authors’ previous work [14], several simulation models for the VRF unit were developed and validated. All the models were
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proved of fast computing and evaporator-number independence, and showed good ability for control analysis. One of the VRF models is used in the simulation of the combined system.
2.3 Building description and air conditioned zone model
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An office building is selected to accommodate the combined system. The building has six
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floors in total above ground. Each floor of the building is divided into six conditioned thermal zones, corresponding to four outdoor exposures (east, west, south and north), an interior zone (including a ring-shaped corridor) and a center core, as shown in Fig. 2. The center core is space for elevators and stairs. The air conditioned zones are numbered from zone 1 to zone 25 for
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conveniently identifying their locations. A summary of the key parameters of the building is listed in Table 1.
By considering each zone as a control volume, ordinary differential equations [31, 32], which
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express the transient behaviors of temperature, humidity and CO2 concentration, are derived and
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described as follows. It is noted here that balances of the sensible load and the latent load are separately considered for having an insight into their variations and in accordance with the modeling methodology of TRNSYS.
M ic p
Mi
Vi
dTi = ms ,VRF ,i c p (Ts ,VRF ,i − Ti ) + ms ,OA,i c p (Ts ,OA − Ti ) + Qi dτ
dwi = ms ,VRF ,i ( ws ,VRF ,i − wi ) + ms ,OA,i ( ws ,OA − wi ) + Di dτ
dCi = vs ,i ( Cs ,i − Ci ) + G ⋅ Occpi dτ
(1)
(2)
(3) 9
ACCEPTED MANUSCRIPT Where, M is the mass of the air, V is the volume of the zone, T is the indoor temperature, ms is the mass flow rate of supply air, Ts is the temperature of supply air, Q is the internal load, w is the indoor humidity ratio, ws is the humidity ratio of supply air, D is the internal humidity load, C is
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the indoor CO2 concentration, Cs is the CO2 concentration of supply air (360 ppm is used), vs is the volume flow rate of supply air, Occp is the number of occupants and G is the amount of CO2
emission rate of people (5×10-6 m3/s is used). The subscript VRF, OA and i represent VRF part,
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OA processing part and ith zone, respectively.
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The internal load (Q) and the internal humidity load (D), generated by occupant, lighting and equipment, are simulated by offering operation schedules to mimic the practical variations. Other factors, including solar radiation, heat transfer through window and envelops are not manually interacted. Table 2 describes the methodology of heat gain calculation for occupant, lighting and
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equipment. LightMax, EqMax and OccMax denote designed lighting load per square, designed equipment load and designed occupancy number, respectively. LSchedule, ESchedule and OccSchedule denote possibility of maximum lighting, number of operating equipment and
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occupancy density, respectively. OHeat denotes heat generation rate per person.
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3. Control performances of the combined system 3.1 Test conditions
Zones 7 to 12 as shown in Fig. 2 are selected for the study because of good representing of internal and external factors such as orientations, radiations, etc. Heat generated by lighting and equipment are designed the same for each zone. Schedules of lighting and equipment are shown in Fig. 3. The profile of occupancy in each zone is not the same and is simply distinguished by multiplying a different occupant factor onto a base occupant density, which is shown in Fig. 3 as 10
ACCEPTED MANUSCRIPT well. The occupant factors for zone 7 to 12 are 0.95, 1.0, 0.8, 1.05, 1.10 and 0.90, respectively. The light schedule (0~1) describes the possibility of maximum lighting to mimic those offices with intensive light that one can turn (some of) them on and off conveniently. The corridor
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adjacent to these zones is assumed to have an independent air conditioner, which has a thermostat set to 24℃ in cooling mode. It is assumed that there is no air exchange between different floors in this paper. 30th July of Shanghai, whose TMY (Typical Meteorological Year) weather profile is
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supposed to be operated during 8 a.m. to 20 p.m.
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shown in Fig. 4, is chosen for representing the typical cooling conditions. The combined system is
3.2 Control performances
Two cases, the same and different temperature set-points of the zones are investigated. For the case of the same set-point, the set-points of all the 6 zones are 24℃, and the OA supply
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temperature set-point is also 24℃. For the case of different set-point, the OA supply temperature set-point is 20℃, zone 7 and zone 12 are 24℃, zones 8 is 25℃, zone 9 and zone 10 are 26℃. In addition, in the case of different set-point, zone 11 is not controlled that the IDU is stopped to test
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how well the combined system adapts to different amount of operating indoor units.
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3.2.1 Control performance of VRF part Figs. 5 and 6 show the control performances of the VRF part in the case of the same set-point and different set-point, respectively. As shown in Fig. 5a and 6a, all the zones can be maintained at their specific set-points, and the control is also stable no matter the set-points change or not. There are tiny peaks and valleys in the temperature profiles of the zones in either of the cases, due mostly to the continuously varying OA flow rate. For zone 11 in the case of different set-point, the temperature changes freely and 11
ACCEPTED MANUSCRIPT rationally due to load variation since it is not controlled. The highest temperature achieves about 30℃ at early afternoon when the cooling load is comparatively large during the day. The zone relative humidity is not directly controlled. Instead, in the process of meeting the sensible load of
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each zone, dew-point temperature of the air is forced below the tube wall temperature, coincidentally providing dehumidification as a result. However, as shown in Fig. 5b and 6b, relative humidity of all the controlled air conditioned zones can meet the thermal comfort
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requirements as they are all maintained at about 40% ~ 65%. The separating of the relative
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humidity in the case of the same set-point is caused by the different occupant condition in each zone. The high temperature and relative humidity of zone 11 in the case of different set-point imply the importance of air conditioning for buildings in summer in Shanghai. Fig. 5c and 5d depict the changing behaviors of the EEVs’ opening and the compressor speed
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(nominated by dividing the rated value, 3000rpm) in the case of the same set-point, respectively. The corresponding variations in the case of different set-point are depicted in Fig. 6c and 6d, respectively. In both cases, all the EEVs except that in zone 11 in the case of different set-point
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quickly respond to the required cooling requirement, which is represented by the difference
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between the zone air return temperatures and their set-points and the change rates of the zone temperatures. In addition, the compressor speed also changes effectively in response to changes in superheated degree caused by variation of EEVs’ opening so that the compressor capacity can closely match the total loads. 3.2.2 Control performance of OA processing part The control loops of the OA processing part are designed for ensuring sufficient ventilation air in order to reduce CO2 concentration, or in other words, to maintain the IAQ. Since CO2 12
ACCEPTED MANUSCRIPT generation is solely related to the number of occupants that it is independent of zone temperature and humidity, OA flow rate as well as CO2 concentration in either of the two cases are the same. Therefore, the control performances in OA processing part will be illustrated here only based upon
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the case of different set-point (Figs. 7 and 8). As is shown in Fig. 7a, the OA flow rate of the combined system varies gradually with the
time. It is adjusted according to the variation of the occupants. The rising up and descending down
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of the OA flow rate and the occupancy are almost synchronized as a result. The valley in early
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afternoon reflects rational absence of the occupants for lunch, rest, etc. The least minimum OA mass flow rate in this research is set for ensuring basic sanitary condition of the floor when occupants are absent.
The OA flow rate is a result of the combined activities of the VAV box control and the supply
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fan control. Fig. 7b and 7c depict the variations of the air damper in each VAV box and the supply fan frequency. The air dampers are all regulated in rational positions to allow appropriate amount of OA to maintain the zone CO2 concentration at or below the limit, 1000 ppm. The supply fan
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frequency is increased or decreased with the openings of the air dampers to maintain the static
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pressure at the set-point, as shown in Fig. 7c. On the other hand, during most of operation time, the openings of the dampers are all smaller than 75%, it means that there is potential for energy-saving optimal control, i.e., the supply fan frequency could be optimized (reduced) to save energy.
The indoor CO2 concentration as shown in Fig. 7d is well controlled at the IAQ-limit during most of the operation time. The reason of the exception of a short time at the beginning can be explained by the hysteresis of the control system. The tiny peaks and valleys of the CO2 13
ACCEPTED MANUSCRIPT concentration profiles are almost synchronized with the OA flow rate, which implies that the more stable the occupancy variation is, the better control of the OA flow rate as well as the CO2 concentration will be.
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Fig. 8 shows the control performances of the DX unit, including the OA supply temperature, opening of the EEV and speed of the compressor. Fig. 8a shows that the OA supply temperature
opening and compressor speed, as shown in Fig. 8b.
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can be quickly and well controlled at the set-point, with the help of quickly adjusting of the EEV
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In conclusion, test results show that the combined system has good ability and performances responding to various conditions. The control strategies are effective and efficiently implemented. In addition, the demand of individual control could also be met by applying this combined system.
4. Energy efficiency characteristics of the combined system
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4.1 Energy consumption under various conditions
Although the OA processing part is designed in basic purpose to supply and process sufficient OA for maintaining IAQ, it has ability to accommodate part of the indoor loads when
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the OA supply temperature is lower enough. In that way, there will be a trade-off between the
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loads taken by the VRF part and by the OA processing part. However, energy consumption of the combined system will not be the same due to varying efficiencies of the individual parts under different load conditions. From the energy saving point of view, the operation strategy that can provide sufficient cooling but consume the least energy will be preferred. Clearly, energy consumption decreases with increasing of COP. Therefore, the energy consumption of the combined system can be reduced or even minimized when both the VRF unit and the OA processing unit operates in the most efficient way. However, it is not easy to make both the VRF 14
ACCEPTED MANUSCRIPT unit and the OA processing unit operate with highest COP simultaneously before the characteristics of energy efficiency are completely understood. To understand the characteristics of energy efficiency of the combined system under different
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conditions, a group of tests (GP-A) are carried out first. Different OA supply temperature set-point, from 18 to 26℃ with a 2℃ increment, is selected for each test, with all the zones’ set-point 24℃. Other testing conditions, including load settings, zone locations, and weather profile, are not
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changed.
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Figs. 9 and 10 show test results of the cooling capacity and energy consumption of the VRF unit and the OA processing unit in GP-A tests, where the cooling capacity and the energy consumption are calculated by integral of the cooling and input power over the operating time, respectively.
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As shown in Fig. 9, the total cooling capacity of the combined system is the same in all the test cases. The cooling capacity of the OA processing part decreases with the increasing set-point of the OA supply temperature, while the cooling capacity of the VRF part acts in a reverse way.
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As shown in Fig. 10, the energy consumption of the VRF unit increases with increasing
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set-point of OA supply temperature for taking more OA loads. However, the energy consumption of the OA processing unit is not monotonely changing with the OA supply temperature. It decreases to the lowest value and then increases with increasing set-point of the OA supply temperature. It means that there is minimum energy consumption for the combined system, i.e., there is a best OA supply temperature that can minimize the energy consumption of the combined system. On the other hand, another group of tests (GP-B) are further carried out with the OA supply 15
ACCEPTED MANUSCRIPT temperature set-point fixed (20℃) and the rated capacity of the OA processing unit (DX unit) varying from 50% to 150% of that used in GP-A in a 25% increment. Other conditions keep the same as GP-A. The results are shown in Fig. 11.
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As shown in Fig. 11, the energy consumption of the OA processing unit decreases to the lowest value when the current DX capacity (i.e. the 100% case) is used. It then increases with increasing DX capacity. It means that the DX capacity affects the energy consumption of the
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combined system. We should carefully select DX capacity for practical applications. Actually, the
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current DX capacity has been determined in GP-A tests for high energy efficiency by trial and error method. However, when the DX capacity is determined, the energy consumption of the combined system is mainly affected by the OA supply temperature, as illustrated in GP-A tests. 4.2 Relationship between COP and PLR
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The interesting results of the total energy consumption in both GP-A and GP-B can be further explained by the relationships between COP and PLR, where they are calculated as follows.
Q Wcom + W fan
PLR =
Q Qrated
(4)
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COP =
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(5)
Where Q and Qrate denote actual and rated cooling of the VRF unit, respectively. W denotes input power, and subscript com and fan denote compressor and supply fan, respectively. It should be noted here that Q and W used here are instantaneous values. Fig. 12 and 13 show the relationships between COP and PLR of the VRF unit and the OA processing unit in GP-A, respectively, where the PLR is calculated as the ratio of the provided cooling to the rated capacity of the individual unit. Generally, for both units, COP decreases with 16
ACCEPTED MANUSCRIPT increment of PLR, and vice versa. When PLR locates in the range from 0.4 to 0.7, COP attains its maximum, which is in accordance with the conclusions of experimental studies [35]. COP falls to the lowest when PLR gets its value larger than 0.8. However, in cases of high OA supply
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temperature set-point (e.g. 26℃), COP of OA processing unit increases or decreases in accordance with PLR strictly, which is caused by the DX compressor having to operate at its minimum value due to relatively small OA load. PLR and COP of the VRF unit change in a narrower range
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comparing to that of the OA processing unit. In other words, OA supply temperature has
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significantly influences on operation of the OA processing unit, and the total energy consumption of the combined system is greatly impacted as a result.
The relationship between COP and PLR of the OA processing unit in GP-B is shown in Fig. 14. That of the VRF unit is not shown again since the operation of the VRF unit is not affected
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when the OA supply condition is fixed. As shown in Fig. 14, COP of the OA processing unit is improved as the DX capacity is decreased during early morning and late afternoon periods when the loads are small. However, the COP is not improved with the decreasing of DX capacity at
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other times. On the other hand, the COP decreases significantly as the DX capacity is increased,
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especially when the capacity is 1.5 times larger. And it is the reason that the energy consumption is increased dramatically in these cases. Specifically, it can be seen from Fig. 13b and 14b that, in most of the operation time, PLR of the OA processing unit locates in the range from 0.4 to 0.7 when the OA supply temperature set-point is 20. So it results in high efficiency operation of the OA processing unit among all the cases. From this point of view, energy consumption of the combined system can be reduced by suitable OA supply temperature set-point, i.e., OA supply temperature set-point could be 17
ACCEPTED MANUSCRIPT optimized according to the operation conditions. In conclusion, results show that the energy efficiency of the VRF unit as well as the OA processing unit changes with variation of load. It reaches the highest when PLR locates in certain
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range, which indicates that further optimization can be realized by making both the VRF unit and the OA processing unit operate in the highest efficiency range.
5. Conclusions and future work
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Simulations of a new air conditioning system combining VRF with OA processing unit in
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cooling conditions are presented. The illustration of the combined system has been introduced and its control loops has been designed. The simulation methodology is developed and the tests are carried out to validate the control performances and energy efficiency characteristics of the combined system.
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The simulation tests results show that all the zones of the combined system could be maintained at their specific set-points within a small error after the control is stable no matter the set-points are the same or different, and no matter the amount of operating indoor units changes or
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not. Besides, the IAQ can be ensured.
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The relationship between COP and PLR indicates that further optimization can be realized by making both the VRF unit and the OA processing unit operate in the high efficiency range. The results also show that there is a best OA supply temperature set-point that can optimize the energy consumption of the combined system. As the OA processing unit affects the operation of the combined system more significantly than the VRF unit, the optimization can be realized by simply controlling the PLR of the OA processing unit to the most efficient system-operation range. In this paper, characteristics of the combined system under regular control strategies are 18
ACCEPTED MANUSCRIPT simulated and analyzed. Preliminary study about optimal control of the combined system is also carried out. Future work could be put forward in several aspects, especially in online optimization
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of the OA supply temperature. The results will be presented in subsequent papers.
Acknowledgements
The research work presented in this paper is financially supported by National Natural Science
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Foundation of China (No. 51376125) and National Natural Science Foundation of China (No.
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51376126).
References
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[2] J.W. Jeong, S.A. Mumma, W.P. Bahnfleth, Energy Conservation Benefits of a Dedicated OA System with Parallel Sensible Cooling by Ceiling Radiant Panels, ASHRAE Trans. 109 (2003) 627-636.
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[3] W. Goetzler, Variable refrigerant flow systems, ASHRAE J. 49 (2007) 24-31.
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[4] Y.P. Zhou, J.Y. Wu, R.Z. Wang, S. Shiochi, Energy simulation in the variable refrigerant flow air-conditioning system under cooling conditions, Energ. Build. 39 (2007) 212-220. [5] Y.M. Li, J.Y. Wu, S. Shiochi, Modeling and energy simulation of the variable refrigerant flow air conditioning system with water-cooled condenser under cooling conditions, Energ. Build. 41 (2009) 949-957. [6] X.B. Liu, T.Z. Hong, Comparison of energy efficiency between variable refrigerant flow systems and ground source heat pump systems, Energ. Build. 42 (2010) 584-589.
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[11] J.M. Choi, Y.C. Kim, Capacity modulation of an inverter-driven multi-air conditioner using electronic expansion valves, Energy 28 (2003) 141-155. [12] W.X. Shi, S.Q. Shao, X.T. Li, X.F. Peng, X.D. Yang, A network model to simulate performance of variable refrigerant volume refrigerant systems, ASHRAE Tran. 109 (2003) 61–68.
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[13] S.Q. Shao, W.X. Shi, X.T. Li, Q.S. Ye, Simulation model for complex refrigeration systems based on two-phase fluid network-Part I: Model development, Int. J. Refrig. 31 (2008)
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ACCEPTED MANUSCRIPT residential use, ASHRAE Trans. 97 (1991) 127-131. [17] W. Chen, X.X. Zhou, S.M. Deng, Development of control method and dynamic model for multi-evaporator air conditioners (MEAC), Energ. Convers. Manag. 46 (2005) 451-465.
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[18] R. Shah, A.C. Alleyne, C.W. Bullard, Dynamic modeling and control of multi-evaporator air conditioning systems, ASHRAE Trans. 110 (2004) 109-119. [19] J.L. Lin, T.J. Yeh, Identification and control of multi-evaporator air conditioning systems, Int. J. Refrigeration 30 (2007) 1374-1385. [20] T.N. Aynur, Y. Hwang, R. Radermacher, Integration of variable refrigerant flow and heat
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[21] T.N. Aynur, Y. Hwang, R. Radermacher, Integration of variable refrigerant flow and heat pump desiccant systems for the cooling season, Appl. Therm. Eng. 30 (2010) 917-927. [22] R. Parameshwaran, R. Karunakaran, C.V.R. Kumar, S. Iniyan, Energy conservative building air conditioning system controlled and optimized using fuzzy-genetic algorithm, Energ. Build.
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ACCEPTED MANUSCRIPT California, USA, pp. 138-147. [27] O.O. Abe, C.J. Simonson, R.W. Besant, W. Shang, Effectiveness of energy wheels from transient measurements. Part I: Prediction of effectiveness and uncertainty, Int. J. Heat Mass
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Transf. 49 (2006) 52-62. [28] T.N. Aynur, Variable refrigerant flow systems: A review, Energ. Build. 42 (2010) 1106-1112. [29] Y.P. Zhou, J.Y. Wu, R.Z. Wang, S. Shiochi, Simulation and experimental validation of the
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(2008) 1041-1047.
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[34] Y. Ye, W. Li, Energy analysis on VAV system with different air-side economizers in China, Energ. Build. 42 (2010) 1220-30. [35] W.H. Xue, P.L. Chen, C.J. Liu, The experimental study of the operating characteristic and energy consumption of the VRV air-conditioning system, Energ. Conserv. Tech. 21 (2003) 3-5. (In Chinese)
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List of Figure captions
Fig. 2 Typical floor plan of the building (unit in drawing: meter) Fig. 3 Base occupant density and schedules of equipment and lighting
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Fig. 4 TMY weather profile of Shanghai in 30th July
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Fig. 1 Schematic and control designing of the combined system
Fig. 5 Control performances of the VRF part in the same set-point case: (a) temperature, (b)
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relative humidity, (c) EEV opening, (d) compressor speed
Fig. 6 Control performances of the VRF part in different set-point case: (a) temperature, (b) relative humidity, (c) EEV opening, (d) compressor speed
Fig. 7 Control performances of OA flow: (a) OA flow rate, (b) air damper opening, (c) supply fan
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frequency, (d) CO2 concentration
Fig. 8 Control performances of DX unit: (a) OA supply temperature, (b) Compressor speed and
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EEV opening
Fig. 9 Cooling capacity of the combined system in GP-A tests
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Fig. 10 Energy consumption of the combined system in GP-A tests Fig. 11 Energy consumption comparison of the combined system in GP-B tests Fig. 12 PLR and COP of VRF unit in GP-A tests Fig. 13 PLR and COP of OA processing unit in GP-A tests Fig. 14 PLR and COP of OA processing unit in GP-B tests
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ACCEPTED MANUSCRIPT Table 1 Critical information of the building Description
Building location Building type and storeys Gross floor area (air conditioned area) Typical floor plan Zone height (m) Solid wood door size Windows and shading
Shanghai, China Office building, 6-story above ground 4700 m2 Area=28 m×28 m, floor-to-floor height=3.5m 2.7 m Width × height =1.0 m×2.0 m Low-e double pane glazing. Window width × height = 1.8 m×1.5 m; sill height=0.80 m; No shading device Aerated concrete wall, thickness=310 mm. Concrete blocks, thickness=200 mm. Reinforced concrete, thickness=120 mm.
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Item
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External wall Internal wall Floor
ACCEPTED MANUSCRIPT Table 2 Settings and calculation method of the heat gain Item
Description
Lighting
Lightload=LightMax × LSchedule × ZoneArea LightMax =20W/m2; LSchedule (0~1). Equipload=EqMax × ESchedule EqMax =120W; ESchedule (0~6). Occpload= OccMax × OccSchedule×OHeat OccMax=6; Seated, very light activity, OHeat=65W.
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> A VRF and ventilation combined system is proposed to take merits of both parts; > A simulation platform is developed for the combined system; > The combined system could maintain thermal comfort and indoor air quality; > Characteristics of system energy efficiency are revealed for further optimal control;