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Performance comparison between bypass cycle and injection cycle for sub-cooling methods in multi-split variable refrigerant flow (VRF) system in hot seasons
Accepted Manuscript
Performance comparison between bypass cycle and injection cycle for sub-cooling methods in multi-split variable refrigerant flow (VRF) system in hot seasons Byungchae Min , Seokhoon Jang , Taemin Lee , Heunghee Bae , Cheoreon Moon , Gyungmin Choi PII: DOI: Reference:
S0140-7007(19)30308-1 https://doi.org/10.1016/j.ijrefrig.2019.07.010 JIJR 4465
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
23 October 2018 26 May 2019 18 July 2019
Please cite this article as: Byungchae Min , Seokhoon Jang , Taemin Lee , Heunghee Bae , Cheoreon Moon , Gyungmin Choi , Performance comparison between bypass cycle and injection cycle for sub-cooling methods in multi-split variable refrigerant flow (VRF) system in hot seasons, International Journal of Refrigeration (2019), doi: https://doi.org/10.1016/j.ijrefrig.2019.07.010
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ACCEPTED MANUSCRIPT Highlights
The bypass cycle and injection cycle in a multi-split VRF system help increase the sub-cooling degree at inlet of EEV.
Energy efficiency ratio (EER) of the multi-split VRF system increases with application of the bypass
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cycle and injection cycle. Performance of the VRF system with the injection cycle is more efficient under the same cooling
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capacity.
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Performance comparison between bypass cycle and injection cycle for sub-cooling methods in multi-split variable refrigerant flow (VRF)
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system in hot seasons
Byungchae Mina, Seokhoon Jangb, Taemin Leea, Heunghee Baeb, Cheoreon Moonb, Gyungmin Choia* a
School of Mechanical Engineering, Pusan National University, Busan 46283,
b
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Republic of Korea
SAC Research/Engineering Division, H&A Solution company, LG Electronics, 84,
*
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Wanam-ro, Seongsan-gu, Changwon-si, Gyeongsangnam-do,51554, Republic of Korea
Corresponding Author: Professor Gyungmin Choi, School of Mechanical
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Engineering, Pusan National University, Busan 46283, Republic of Korea Telephone: +82-51-510-2476, Fax: +82-51-512-5236
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Email: [email protected]
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Abstract
In this study, the performance of a multi-split VRF system using the bypass cycle and injection cycle is
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evaluated using the numerical simulation as a possible sub-cooling method to prevent flash gas generation in liquid pipelines. The simulation for the multi-split VRF system is developed by considering the
applications of the bypass cycle and injection cycle, and is validated with experimental data. The bypass cycle and injection cycle in the multi-split VRF system yield improvements in their cooling capacities of the order of 3.22 % and 13.43 %, respectively, and energy efficient ratio (EER) of the order of 1.98 % and 1.72 %, respectively. The input power of the injection cycle is reduced by up to
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ACCEPTED MANUSCRIPT 4.45 % when the performance of the multi-split VRF systems with bypass cycle and injection cycle is compared under the same cooling capacity conditions.
Keywords: Variable refrigerant flow system, Sub-cooling method, Bypass cycle, Injection cycle, Numerical
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simulation
NOMENCLATURE : Area (m2)
C
: Flow coefficient (-)
COP
: Coefficient of Performance (-)
h
: Enthalpy (kJ kg-1 )
HT
: Heat transfer
k
: Ratio of specific heat (-)
m, M
: Mass (kg)
ṁ
: Mass flow rate (kg s-1)
P
: Pressure (kPa)
Q
: Capacity (kW) and heat transfer rate (kW)
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A
Rbyp
: Bypass ratio (-)
rcr
: Critical pressure ratio (-)
Rinj
: Injection ratio (-)
T
: Temperature (K)
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ACCEPTED MANUSCRIPT V
: Volume (m3)
v
: Specific volume (m3 kg-1)
W
: Input power (kW)
: Efficiency (%)
θ
: Angle (degree)
ρ
: Density (kg m-3)
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η
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Greek letter
Subscript : average
byp
: bypass
cal
: calculation
comp
: compressor
cond
: condenser
d
: downstream
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M
avg
dis
: discharge
EEV
: electronic expansion valve
eva
: evaporator
g
: gas-phase
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ACCEPTED MANUSCRIPT : internal heat exchanger, sub-cooler
in
: inlet
ind
: indicated
inj
: injection
init
: initial
l
: liquid-phase
m
: mass
mech
: mechanical
motor
: motor
out
: outlet
pipe
: pipeline
sc
: sub-cooling
set
: setting
sh
: super-heating
suc
: suction
u
: upstream
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AC
vi
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IHX
: vapor-injection
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ACCEPTED MANUSCRIPT Introduction The multi-split variable refrigerant flow (VRF) system has been used to serve cooling or heating units in the middle or large-sized buildings, such as residential and commercial buildings (ex. apartments, hotels, theaters, gyms, etc.), because this system can adjust the cooling or heating capacity for each separated zone according to the cooling or heating loads using an inverter scroll compressor and electronic expansion valves (EEV). In
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addition, the multi-split VRF system consists of at least one outdoor unit and several indoor units, which have different capacities and types, thereby reducing the installation space of outdoor units at the ground or rooftop of the buildings. However, pipelines between outdoor and indoor units can be lengthened by at most 150 m among existing VRF system products (Afify, 2008). Because the multi-split VRF system is used in the middle
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or large-sized buildings with long pipelines, the system should be designed by considering the pressure drop in the pipelines between the outdoor and indoor units since the pressure drop causes performance degradation of the overall system owing to the increased energy consumption required to transfer the working fluid and the flash gas generation in the liquid pipelines. Specifically, the flash gas generation in the liquid pipeline leads to
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noise in EEVs, decrease of the mass flow rate through the EEV, and yields lower system efficiency and reliability. Kang et al. evaluated the influence of flash gas generation in a refrigeration system. As the result of
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their study, the mass flow rate through an EEV decreased due to the increase of the void fraction with the increase of flash gas ratio. Hence, the cooling capacity and input power decreased with the increase of flash
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gas ratio because of reduced mass flow rate, and the coefficient of performance (COP) of the refrigeration system was degraded significantly since a decrease of the cooling capacity was larger than that of the input
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power with an increase of flash gas ratio (Kang et al., 2012). Therefore, in the multi-split VRF system, a sub-
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cooler heat exchanger is typically used to prevent flash gas generation in the liquid pipeline. In a multi-split VRF system, two types of refrigeration cycles that use the sub-cooler to prevent flash gas generation exist. The first one is a bypass cycle, and the other is an injection cycle. Fig. 1 shows a schematic of the bypass cycle and injection cycle. In the bypass cycle, the bled refrigerant is expanded at a low temperature, which is similar to the response of the evaporating temperature in the system. The low-temperature refrigerant in the bypass stream is heated by the heat exchange with a high temperature in the main stream, and the bled refrigerant is finally mixed with the refrigerant flowing out from the evaporator in the accumulator. In the injection cycle, the bled refrigerant
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ACCEPTED MANUSCRIPT is expanded to a higher pressure than the bypass pressure. The bled refrigerant is finally injected into a compression chamber in a scroll compressor. In both cycles, a larger sub-cooling degree of the system can be elicited with the sub-cooler compared to the system without the sub-cooler, thereby avoiding flash gas generation and increasing the cooling capacity. Moreover, when the bled refrigerant is in two-phase state and temperature of the bled refrigerant is not higher than that at the suction port of a compressor, the discharge
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temperature of a compressor can decrease due to a decrease of suction temperature of a compressor in the bypass cycle. In addition, the discharge temperature can decrease by an intercooling effect of the refrigerant injection in the injection cycle.
There are some studies that investigated the performance of a multi-split VRF system by using a sub-cooler
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heat exchanger with the bypass cycle and injection cycle in the literature. Kwon et al. investigated the effect of the sub-cooler heat exchanger with a bypass cycle on the performance of a multi-split VRF system, based on field performance tests in educational offices during the cooling season. In their study, it was revealed that a multi-split VRF system with a sub-cooler heat exchanger improved the cooling performance factor (CPF) by
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approximately 8.5 % compared to a system without a sub-cooler heat exchanger (Kwon et al., 2012). Li et al. analyzed the effect of a sub-cooler heat exchange in a multi-split VRF system with the bypass cycle by
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varying three factors, namely, the length of the pipeline, the capacity of the sub-cooler heat exchanger, and the bypass refrigerant mass flow rate ratio. Their simulation results indicated that the COP of a multi-split VRF
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system can be improved with a long pipeline by controlling the appropriate bypass mass flow rate ratio (Li et al., 2016). Tu et al. carried out experiments for the analyses of the effects of the sub-cooler on the cooling
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capacity, energy efficiency ratio, degree of sub-cooling, and degree of discharge superheating in a four-indoor
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VRF system with the bypass cycle. In accordance to their experimental results, they found that the reasonable adjustment of opening of the sub-cooler EEV can improve the performance of a multi-split VRF system by adopting the bypass cycle. The maximum energy efficiency ratio of the system with a sub-cooler was improved by up to 3.3 %, 3.1 %, and 3.0 %, when one, two, and four indoor units were operated on the thermo-on mode, respectively (Tu et al. 2017) . Wang et al. adopted an internal heat exchanger (sub-cooler) in a single-unit heat pump system with an injection cycle to enhance the performance of the system in hot and cold climates. Vapor injection in a scroll compressor improved the performance of the heat pump system because the heat exchange between the main and injected streams in the sub-cooler reduced the refrigerant
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ACCEPTED MANUSCRIPT enthalpy at the inlet of the evaporator with an increase of the degree of sub-cooling (Wang et al. 2009). Cho et al. measured and compared the cooling and heating capacities of R410A and R32 multi-heat pump systems with a vapor injection. The cooling and heating capacity of the system with R410A and R32 were enhanced by vapor injection owing to an increase of the sub-cooling degree at the inlet of the EEV using a sub-cooler (Cho et al., 2016).
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In the studies mentioned previously, it was found that use of a sub-cooler can increase the degree of subcooling for both bypass cycle and injection cycle. Therefore, flash gas generation in the liquid pipeline can be prevented by the sub-cooler. The bypass cycle and injection cycle can be applied to the same multi-split VRF system with a sub-cooler by using solenoid valves to switch the bled stream path into the accumulator or into
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the compression chamber. In addition, it can be found that the multi-split VRF system using the bypass cycle and injection cycle has different operating characteristics in the literature. However, the study on performance comparison between both cycles in the multi-split VRF system based on the numerical simulation cannot be found in the literature.
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The objective of this study is to compare the operating characteristics of both multi-split VRF systems using the bypass cycle and injection cycle based on the numerical simulation. The bypass and injection ratios are
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important factors to determine the performance of a multi-split VRF system. Therefore, the operating characteristics of both systems are compared by varying the bypass and injection ratios in the same
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configuration of the multi-split VRF system and by using the same operating conditions. Moreover, the improvement potential of two cycles in the multi-split VRF system is compared with optimum bypass ratio
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and injection ratio under same cooling capacity.
2. Mathematical models 2.1 Scroll compressor To simulate the performance of a compressor, three models were considered, such as the semi-empirical model (Sun et al., 2017, Zhu et al., 2013, Shao et al., 2012), map-based model (10-efficient ARI compressor map), and the geometry-based model (Kwon et al., 2017).The semi-empirical and map-based models have been used in most previous studies because these models are simpler in numerical simulations with experimental data to reduce the simulation time. However, these models are not suitable for the simulations of
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ACCEPTED MANUSCRIPT heat pumps equipped with a refrigerant-injection scroll compressor. This is because the refrigerant injection into a compression chamber in a scroll compressor occurs continually during the compression process, as shown in Fig. 2. This depends on many parameters, including the injection pressure, suction pressure, position of an injection hole, sub-cooler type and capacity, and others. Hence, Kwon et al. adopted the geometry-based thermodynamic model in the simulation of a vapor injection heat pump for electric vehicles in order to
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consider the vapor injection process in a scroll compressor. In the bypass cycle, the degree of superheating at compressor suction, which significantly affects the input power and discharged mass flow rate of a scroll compressor, is changed with the variation in the bypass mass flow rate ratio. Moreover, in the injection cycle, the input power and discharged mass flow rate of the scroll
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compressor are significantly affected by the injection process. In this study, therefore, the geometry-based thermodynamic model was applied to the simulation of the multi-split VRF system with a sub-cooler in order to take into account the refrigerant injection process and degree of superheating at compressor suction. The governing equations for the numerical model of a scroll compressor include changes in mass, enthalpy,
=
𝑑𝑚𝑖𝑛 𝑑𝜃
−
𝑑𝑚𝑜𝑢𝑡 𝑑𝜃
𝑑ℎ
+
𝑑𝑚𝑖𝑛𝑗 𝑑𝜃
(1)
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𝑑𝑚 𝑑𝜃
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pressure, and temperature in a pocket, as respectively indicated by Eqs. (1)–(2) (JSRAE, 2018).
𝑑𝑃
(2)
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M 𝑑𝑡 = 𝑚̇𝑖𝑛 (ℎ𝑖𝑛 − ℎ) − 𝑚̇𝑜𝑢𝑡 (ℎ𝑜𝑢𝑡 − ℎ) + 𝑚̇𝑖𝑛𝑗 (ℎ𝑖𝑛𝑗 − ℎ) + 𝑄 + 𝑉 𝑑𝑡
where Q is heat transfer rate between the wrap wall and the refrigerant in a pocket. The pressure and
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temperature in a control volume are calculated using the density and enthalpy obtained from Eqs. (1)-(2). REFPROP 9.0 was used to calculate the properties of the working fluids. In this study, the governing equations for the numerical model of the scroll compressor are formulated based on the following assumptions: a) Leakages are ignored between pockets b) Heat transfer and oil effects in a scroll compressor are ignored. Therefore, Q in Eq. (2) is zero kW c) Properties in a volume are homogeneous
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ACCEPTED MANUSCRIPT d) The gravity and kinetic energy are ignored Mass flow rates in a scroll compressor, such as suction, injection, and discharge flows, are predicted by Eqs. (3)–(5) (JSRAE, 2018)..
2/𝑘
2𝑘 𝑃 𝑚̇ = 𝐶𝐴√((𝑘−1) 𝑃𝑢 𝜌𝑢 ) {( 𝑃𝑑 )
−
}
(3)
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𝑢
𝑘+1
𝑃 𝑘 ( 𝑃𝑑 ) 𝑢
𝑘
𝑟𝑐𝑟 =
2 𝑘−1 (𝑘+1)
when
𝑃𝑑 𝑃𝑢
(4)
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< 𝑟𝑐𝑟 , then 2
𝑚̇ =
2𝑘 2 𝑘−1 𝐶𝐴√𝑘+1 𝑃𝑢 𝜌𝑢 {(𝑘+1) }
(5)
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When the refrigerant injection flow is two-phase flow (𝑥𝑖𝑛𝑗 < 1), the upstream density (𝜌𝑢 ) is calculated by
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Eq. (6) (JSRAE, 2018).
(6)
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𝜌𝑢 = 𝜌𝑙 + 𝑥(𝜌𝑔 − 𝜌𝑙 )
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Eq. (6) is a homogeneous model for two-phase flow. The injection mass flow rate is predicted by the sum of the injection mass flow rate at every crank angle. Furthermore, the bypass and injection ratios are defined as
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follows:
𝑅𝑏𝑦𝑝 =
𝑅𝑖𝑛𝑗 =
𝑚̇𝑏𝑦𝑝
(7)
𝑚̇𝑑𝑖𝑠
𝑚̇𝑖𝑛𝑗
(8)
𝑚̇𝑑𝑖𝑠
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ACCEPTED MANUSCRIPT The bypass ratio is the ratio of the discharged mass flow rate to the mass flow rate bypassed into the bleeding line. The injection ratio is the ratio of the discharged mass flow rate to the mass flow rate injected into a compression chamber. In the bypass cycle, summation of the mass flow rates bypassed and flowing in the evaporator is same with the suction mass flow rate which is same with the discharged mass flow rate, while in the injection cycle, the discharged mass flow rate is equal to summation of the suction and injection mass flow
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rates. The indicated work of a scroll compressor can be predicted by the variation in pressures and volumes of the pockets from the suction to the discharge processes, as indicated by Eq. (9).
𝑊𝑖𝑛𝑑 = ∮ 𝑃 𝑑𝑉
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(9)
Finally, the compressor input work is calculated using Eq. (10), including the motor and mechanical efficiencies. The motor, mechanical, and volumetric efficiencies are obtained from compressor calorimeter
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tests with variation in the operating conditions such as suction and discharge pressures, and rotational speed.
(10)
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𝑊𝑐𝑜𝑚𝑝 = 𝑊𝑖𝑛𝑑 /𝜂𝑚𝑜𝑡𝑜𝑟 𝜂𝑚𝑒𝑐ℎ
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The data of these efficiencies are formulated by the linear regression.
The injection flow area is calculated using an equation derived by Liu et al. in relation to the orbiting angle
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the involute.
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(Liu et al., 2009). The injection hole was located at 300° in the inward direction from the suction port along
2.2 Heat exchangers 2.2.1 Condenser and evaporator The multi-split VRF system adopted a louver fin-tube heat exchanger in an outdoor unit as a condenser when operated in the cooling mode. In this study, the tube-by-tube method is applied to the heat exchanger
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ACCEPTED MANUSCRIPT simulation to calculate the heat transfer between the refrigerant and air, and the pressure drop of the refrigerant and air sides in one tube defined as a control volume with the effectiveness–NTU method.
2.2.2 Sub-cooler A tube-in-tube heat exchanger type, which consists of one outer tube and seven inner tubes, is used as the sub-
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cooler. The flow area of the outer tube is obtained by the equivalent diameter to calculate the flow velocity, Reynolds number, and the heat transfer coefficient. The heat transfer in the sub-cooler is calculated by adopting the effectiveness-NTU method.
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2.3 Pipelines
The liquid pipeline and the gas pipeline are connected between the outdoor and indoor units. In the multi-split VRF system, the system performance is affected by the pressure drop and heat transfer in these pipelines. Hence, in this study, the pressure drop and the heat transfer in the pipelines are modeled to consider the effects
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of these factors on the overall system performance. The correlations used for the model of the heat exchangers
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and the pipelines are summarized in Table 1.
2.4 Electronic expansion valve (EEV)
= 𝐶𝑣 𝐴√𝜌
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𝑚̇
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The mass flow rate flowing through an EEV can be predicted by the orifice equation,
,𝑖𝑛
where A is the flow area,
𝑃
𝑃
(11)
is the pressure difference between the inlet and outlet of the EEV, 𝜌
,𝑖
is the
refrigerant density at the inlet, and 𝐶𝑣 is the flow coefficient which is obtained from the literature (Zhu et al., 2013). 2.5 Multi-split VRF system equipped with a sub-cooler In the simulation of a multi-split VRF system, the basic laws of mass, energy, and momentum conservation of the refrigerant properties in the entire system should be satisfied, as indicated by Eqs. (12) - (14).
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𝑚̇𝑐𝑜𝑚𝑝 − ∑ 𝑚̇
− 𝑚̇𝑖𝑛𝑗 = 0
(12)
𝑊𝑖𝑛,𝑐𝑜𝑚𝑝 + ∑𝑛𝑖=1 𝑄𝑒𝑣𝑎,𝑛 - 𝑄𝑐𝑜𝑛𝑑 = 0
(13)
𝑃𝑑𝑖𝑠 − 𝑃𝑠𝑢𝑐 =
(14)
+ 𝑃𝑝𝑖𝑝𝑒
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𝑃
Fig. 3 shows the flow chart of the simulation logic for the multi-split VRF system which can satisfy the governing equations of the entire system. As shown in Fig. 3, the simulation logic consists of five sub-loops
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for convergence based on the set criterion with the use of an iterative scheme. In the first loop, mass flow rate from the compressor and discharge enthalpy are predicted by the compressor model. The discharge enthalpy is used as the enthalpy at the inlet of the condenser with assuming that the heat transfer and pressure drop in the pipe between the compressor and condenser are neglected. Subsequently, the enthalpy at the outlet of the
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condenser is calculated by the condenser model. In order to satisfy the preset sub-cooling degree at the outlet of the condenser, the calculations of the compressor and condenser models are repeated by adjusting the
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condensing pressure. The second loop is for the convergence of the degree of superheating at exit of injection line, when the degree of superheating at exit of injection line is given as the input data. In this loop, the
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pressure drop and heat transfer in the liquid pipeline were predicted by the pipeline model. In the third loop,
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the mass flow rate was calculated through an EEV in each indoor unit. When the mass balance between the summation of the mass flow rate for all EEVs and the mass flow rates in the evaporators is not satisfied, the
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evaporating pressure is adjusted to match the summation of the mass flow rate for all EEVs with that in the evaporators. At the fourth loop, the cooling capacity is predicted based on the evaporator model using the inlet conditions of the evaporators obtained at the previous steps and assuming an evaporating pressure. If the superheating degree at the outlet of the evaporators is not matched with the set value, the opening areas of the EEVs are adjusted according to the current superheating degree. At the last loop, the pressure drop and heat transfer in the gas pipeline are calculated to obtain the degree of superheating at compressor suction. If the
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ACCEPTED MANUSCRIPT calculated suction enthalpy of a compressor is not the same as the initially assumed value, the assumed suction enthalpy is substituted by the calculated value. The iterative calculation is repeated until the relative and absolute errors are less than their acceptable values. The relative and absolute errors are defined in accordance to Eqs. (15)–(18).
𝑠𝑐,𝑎𝑣𝑔
−
𝑠𝑐,𝑠𝑒𝑡 |
0.1
𝑒𝑠𝑢𝑝,𝑒𝑣𝑎 = |
𝑠ℎ,𝑒𝑣𝑔
−
𝑠ℎ,𝑠𝑒𝑡 |
0.1
𝑒𝑠𝑢𝑝,
= |
𝑠ℎ,𝑖𝑛𝑗
−
𝑠ℎ,𝑖𝑛𝑗,𝑠𝑒𝑡 |
0.1
𝑒𝑚
=
| ∑ 𝑚̇
−𝑚̇ 𝑣 |
(15)
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𝑒𝑠𝑢𝑏,𝑐𝑜𝑛𝑑 = |
(16) (17)
0.1%
(18)
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𝑚̇ 𝑣
3. Experimental apparatus
The specifications for the object system are summarized in Table 2. The rated cooling capacity of the object
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VRF system in this study is 22.4 kW, which comprises one outdoor unit and nine indoor units. Fig. 4 shows the schematic of the experimental apparatus for the multi-split VRF system. The outdoor unit consists of a
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scroll compressor with the displacement volume of 48.0 cc 𝑟𝑒𝑣 −1, a fin-and-tube heat exchanger of the louver-fin type, sub-cooler with the tube-in-tube type, a receiver, an accumulator, EEVs, and a 4-way valve. In
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addition, solenoid valves are used in order to switch the injection cycle mode or the bypass cycle mode, or
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close the paths for the injection line and the bypass line. In order to store the remainder of refrigerant, a receiver is installed at outlet of the condenser in the outdoor unit.
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The indoor unit consists of a heat exchanger of the louver-fin type and an EEV to control the mass flow rate with respect to the degree of superheating at exit of the evaporator. Cooling or heating capacities of the multi-split VRF system are measured in the environmental chambers which can maintain the ambient temperature and humidity, and are composed with one outdoor chamber and two indoor chambers as shown in Fig. 4. The air flow rates of the indoor units are measured by a differential pressure transmitter across the standard nozzle. Power consumption of the outdoor unit and indoor units is measured by a power meter. The suction and discharge pressures are measured by a pressure transducer. Dry
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ACCEPTED MANUSCRIPT bulb and wet bulb temperatures are measured by RTD sensors. The uncertainties for the measurement devices are shown in Table 3. Measurement of the cooling and heating capacities for an indoor unit is repeated six times with three samples of the indoor unit. The repeatability is also shown in Table 3. The experiment on the performance of the multi-split VRF system is carried out in accordance with the AHRI Standard 1230. When the cooling performance was measured, the refrigerant injection technique was not uses
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in the VRF system, but it was used when the experiment on the heating performance was conducted. The experimental conditions according to the AHRI Standard 1230 are summarized in Table 4.
4. Results and discussion
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4.1 Validation of the simulation
The simulation results are compared with experimental data in order to confirm the reliability of the simulation developed in this study. First, the simulation of the scroll compressor is validated based on the compressor calorimeter test. However, in the calorimeter test, details on the injection process could not be
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measured. Therefore, the calculated mass flow rate and input work for the scroll compressor without the refrigerant injection are compared with the calorimeter test data. Figs. 5 (a) and (b) show the validation results
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for the scroll compressor without the refrigerant injection. As shown in Fig. 5, the simulation results for the input work and the mass flow rate of the scroll compressor without the refrigerant injection show good
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agreement with the results of the calorimeter test, and are within 15 % and 5 %, respectively. The simulation results for the multi-split VRF system in the cooling mode are compared to the experimental
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data obtained without the refrigerant injection. Fig. 6 (a) shows the validation results for the simulation of the
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multi-split VRF system in the cooling mode with variation in the part load. As shown in Fig. 5 (a), the simulation results for the cooling capacity and input power of the multi-split VRF system in cooling mode are consistent with experimental data within 15 %. Despite of validation of the system simulation in the cooling mode, it is also necessary to check the accuracy of the simulation for the multi-split VRF system to which the refrigerant injection is applied. The refrigerant injection in a scroll compressor was applied to the experiment for the multi-split VRF system in the heating mode. The validation results for the simulation of the system with an injection cycle in heating mode are indicated in Fig. 6 (b) and show good agreement with a maximum
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ACCEPTED MANUSCRIPT error of 10.3 %. As the results of the simulation, the injection ratios for two-operating conditions in heating mode are 11.6 % and 19.7 %, respectively.
4.2 Performance comparison of a multi-split VRF system using bypass cycle and injection cycle as subcooling methods
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In this study, the performance of a multi-split VRF system is compared according to the variations of the bypass and injection ratios when the sub-cooling methods are employed in the same configuration (length of pipes, capacity of sub-cooler, etc., as shown in Table 2), and the same operating conditions shown in Table 5 in the cooling season. The sub-cooling degree at the outlet of the condenser and the degree of superheating at
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the outlet of the evaporators are fixed as the input date in all cases of the simulation as shown in Table 5. In addition, the simulation was performed for two cases of ambient temperature according to the ASHRAE Standard 37 (ANSI/ASHREA Standard 37-2005, 2005).
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4.2.1 Sub-cooling degree at the inlet of EEV and degree of superheating
The sub-cooling degree at the inlet of the EEV and the degree of superheating at the exit of the bleeding line
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for the bypass cycle and injection cycle are considerably influenced by the bypass and injection ratios, as shown in Figs. 7 (a) and 7 (b) . The sub-cooling degree at the inlet of the EEV increases as the bypass and
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injection ratios increases, but these values become almost constant at the specific value of the bypass and injection ratios for both cycles in the two ambient temperature conditions, as shown in Fig. 7 (a). This is
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because the two-phase heat transfer area in the sub-cooler increases as the degree of superheating at the exit of
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the bleeding line decreases with increases of the bypass and injection ratios. The increase of the two-phase heat transfer area in the sub-cooler leads to the enhancement of the overall heat transfer coefficient according to the increase of the bypass and injection ratios. The heat transfer coefficient in the two-phase region is significantly higher than that in the single-phase. Therefore, the overall heat transfer coefficient is enhanced with the increase of two-phase heat transfer area, but it is not significantly changed after the degree of superheating at the exit of the bleeding line reaches 0 °C because the heat transfer in the sub-cooler occurs just in the two-phase region of the bleeding line without any more increase of two-phase heat transfer area, thereby leading to constant sub-cooling degree.
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ACCEPTED MANUSCRIPT In Fig. 7 (a), the sub-cooling degree of the bypass cycle is higher than that of the injection cycle because of lower temperature in the bleeding line and less mass flow rate in main line compared to the injection cycle. The bled refrigerant in the bypass line is expanded near the evaporating pressure. However, the injection pressure is higher than the evaporating pressure to inject the refrigerant into the compression chamber during the compression process. Hence, in the bypass cycle, the temperature difference in the sub-cooler is larger
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than that in the injection cycle. In addition, the mass flow rates in the main and bleeding lines in case of the bypass cycle are less than those in case of the injection cycle as shown in Fig. 8. Therefore, the sub-cooling degree at inlet of EEV and the degree of superheating at exit of bleeding line in the bypass cycle are higher compared to the injection cycle as shown in Fig. 7.
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As shown in Fig. 7 (a), in case of the baseline with the bypass and injection ratios of 0 %, the sub-cooling degree at the inlet of EEV is lower than the setting value at exit of the condenser with 2.0 °C owing to heat gain from the ambient and pressure drop in the liquid pipeline. In general, the heat gain and pressure drop in the liquid pipeline causes the flash gas generation at EEV with low sub-cooling degree. However, in both
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cycles, despite the heat gain and pressure drop, the sub-cooling degree at the inlet of EEV is higher than that at the exit of the condenser. Therefore, it is possible to prevent flash gas generation in the liquid pipeline with
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appropriate bypass and injection ratios using the sub-cooler in the both cycles.
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As shown in Fig. 7 (b), the bypass ratio shows also an important influence on the degree of superheating at compressor suction (point 1 in Fig. 1) for the bypass cycle, but the degree of superheating at compressor
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suction is not influenced by the variation in the injection ratio. In the bypass cycle, the bled refrigerant is mixed with refrigerant flowing out from the evaporators in the accumulator after the completion of the heat
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exchange with the refrigerant in the main stream. Hence, the degree of superheating at compressor suction is affected by the degree of superheating at the exit of the bypass line. However, in the injection cycle, the degree of superheating at compressor suction is constant despite the variation in the injection ratio because the bled refrigerant is injected directly into a compression chamber. The constant degree of superheating at compressor suction is because of using the liquid receiver tank in the multi-split VRF system as shown in Fig. 4. This receiver tank is designed to hold the liquid refrigerant in order to vary the refrigerant flow rate in the overall system. In the experiment on the multi-split VRF systems, the refrigerant flow rate in the evaporator
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ACCEPTED MANUSCRIPT can be maintained by the store of the refrigerant in the receiver tank despite of the increase of injection ratio. In the simulation, it is assumed that the refrigerant charge is sufficient to supply it into the evaporators. In the bypass cycle, the receiver tank is also used. However, the suction mass flow rate is determined by the compressor displacement volume, rotational speed, and suction density. Therefore, the mass flow rate in the evaporators decreases with the increase of the bypass ratio.
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Fig. 8 shows the total mass flow rate discharged from the compressor and the mass flow rate in the evaporators. In the case of the bypass cycle, the total mass flow rate is not changed significantly at the beginning of the bypass, but increases slightly with additional increases of the bypass ratio because of the decrease of the degree of superheating at compressor suction. The mass flow rate in the evaporators decreases
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continually with the increase in the bypass ratio because of the increase of the bypassed mass flow rate. In the case of the injection cycle, the total mass flow rate increases continually with the increase of injection ratio because the injected refrigerant is added to the suction mass flow rate. The evaporating pressure is not affected significantly by the injection ratio. Therefore, the variation in the suction density of the compressor is very
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small.
Fig. 9 shows the discharge temperature for the two cycles and the two ambient temperature conditions as a
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function of the bypass and injection ratios. The discharge temperature slightly increases at the beginning of the bypass and injection for both cycles at the two ambient temperature conditions, but the temperature starts
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to decrease as the bypass and injection ratios increase further. The reason for the decrease of the discharge temperatures is different for each cycle. In the bypass cycle, the discharge temperature is influenced by the
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suction temperature. However, in the injection cycle, it is influenced by the temperature of the injected
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refrigerant owing to the intercooling effect. Therefore, in the bypass cycle, the discharge temperature increases at the beginning of the bypass because of increase of the degree of superheating at compressor suction, but the discharge temperature decreases as the bypass ratio increases further owing to decrease of the degree of superheating at compressor suction as shown in Fig. 7 (b). In the injection cycle, the discharge temperature also increases at the beginning of the injection owing to high temperature of the injected refrigerant. However, the discharge temperature decreases as the injection ratio increases further because the intercooling effects in the compression chamber is enhanced owing to the lower temperature of the injected refrigerant and the increase of injection mass flow rate. When the ambient temperature is high as the case of 46.1 °C, the
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4.2.2 Cooling capacity and input power Fig. 10 (a) shows the cooling capacities for both cycles at two ambient temperature conditions. The variation trends for both cycles are different as the bypass and injection ratios increase. In the bypass cycle, the cooling capacity increases with the increase of the bypass ratio. However, it decreases from the specific bypass ratio in
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the two cases of different ambient temperatures. This is because the mass flow rate in the evaporators decreases continually despite the increase of the sub-cooling degree with the increase of the bypass ratio. In the injection cycle, the cooling capacity increases continually with the increase of the injection ratio for the two cases of ambient temperature since the mass flow rate in the evaporators are almost constant despite the
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increase of the injection ratio, while the sub-cooling degree increases as the injection ratio increases. The cooling capacity in the injection cycle increases rapidly after the two-phase injection starts because the density
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of the two-phase state is considerably larger than that of the vapor state. The input power is also influenced by the bypass and injection ratios as shown in Fig. 10 (b). In the case of the
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bypass cycle, the input power increases slightly with an increase of the bypass ratio because of the increase of the suction temperature of the compressor and the total mass flow rate, but it decreases as the bypass ratio
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increases further owing to the decrease of the suction temperature. In contrast, the input power in the injection
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cycle increases continually as the injection ratio increases because of the continuous increase of the total mass flow rate. The input power for the injection cycle also increases rapidly after the two-phase injection starts. The cooling capacities in both cycles decrease with the increase of the ambient temperature because the enthalpy at the inlet of evaporator increases owing to the high condensing saturated temperature. The input powers increase also in the high ambient temperature because of higher compression ratio. Fig. 10 (c) shows energy efficiency ratio (EER) as a function of the bypass and injection ratios. EER for the bypass cycle increases gradually with the bypass ratio, but the EER decreases from the specific bypass ratio because the cooling capacity in the two cases of different temperatures decreases despite of reduction of the
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ACCEPTED MANUSCRIPT input power. EER increases at the last point despite of the decrease of cooling capacity because the compression work decreases as the degree of superheating at compressor suction is reached at 0 °C. The increase of EER is outstanding in case of 46.1 °C. In the injection cycle, the variation trend is different from the bypass cycle. At the beginning of the refrigerant injection, EER of the injection cycle decreases because the degree of superheating for the injected refrigerant is high, thereby increasing the specific volume of the
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injected refrigerant. The compressor consumes more compression work with larger specific volume of the refrigerant vapor. However, EER of the injection cycle slightly increases when the degree of superheating at the injection line is reached to 0 °C because of the decrease of specific volume for the injected refrigerant. However, EER of the injection cycle decreases again because the compression work increases sharply with
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starting of the two-phase injection. The highest EER for the bypass cycle in two ambient temperature conditions is attained at the bypass ratio of 15.0 % because the cooling capacities which are influenced by the mass flow rate are highest at this bypass ratio. The EER for the bypass cycle is enhanced up to 1.19 % and 1.98 %, respectively in two ambient temperatures compared to the baseline cycle with the bypass ratio of 0 %.
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In case of the injection cycle, the EER is highest when the degree of superheating of injection line is close to 0 °C, which corresponds to the injection ratios of 10.67 % and 16.01 %, respectively in two ambient
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temperatures. The EER for the injection cycle is enhanced up to 0.42 % and 1.72 % with those injection ratios compared to the baseline cycle with the injection ratio of 0 %. The EER of the bypass cycle is higher than that
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of the injection cycle. However, the cooling capacity for the injection cycle is considerably higher than that of
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the bypass cycle because of constant mass flow rate and the increase of enthalpy difference in evaporators.
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4.3 Performance comparison for the bypass cycle and injection cycle with the same cooling capacity In this section, the performance for both cycles is compared for the same cooling capacity condition at two different ambient temperature conditions. As shown in Fig. 10 (a), the cooling capacity of the injection cycle is higher than the maximum capacity of the bypass cycle. For the comparison between the bypass cycle and injection cycle with the same capacity, the cooling capacity of the injection cycle is matched to the maximum capacity of the bypass cycle because EER of the bypass cycle is higher than that of the injection cycle. The maximum capacities of the bypass cycle for 35.0 °C and 46.1 °C are reached with the bypass ratios of 15.0 %. In order to achieve those maximum capacities in the injection cycle the rotational speed of the compressor is
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ACCEPTED MANUSCRIPT adjusted. In addition, the degree of superheating at the exit of the injection line is set at 2.0 °C, thus allowing the adjustment of the injection pressure based on the iterative method in accordance to the simulation logic, as indicated in Fig. 3. Figs. 11 (a) and (b) show the comparison results for both cycles at the same cooling capacity condition for two ambient temperatures. The same capacities for the injection cycle are attained at the injection ratios of 10.71 %
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and 15.23 %, respectively. The input power of the injection cycle is reduced compared to that of the bypass cycle for 35.0 °C and 46.1 °C, which corresponds to 2.32 % and 4.45 %, respectively. This is because the rotational speed of the compressor could be reduced in order to attain the same cooling capacity as the maximum capacity achieved in the bypass cycle with the compressor rotational speed of 60 Hz and 57 Hz. As
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the input power is reduced in the injection cycle, the EER is also enhanced at the two ambient temperatures by up to 2.33 % and 5.08 %, respectively. In addition, the injection cycle is more effective in decreasing the discharge temperature at high-ambient temperatures, as shown in Fig. 11 (b). Specifically, the discharge temperature is reduced considerably in the injection cycle at the high-ambient temperature condition with the
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same cooling capacity. This is because that in the injection cycles, the same capacities for 35.0 °C and 46.1 °C are attained with the injection ratios of 10.7 % and 15.2 %, respectively. The higher injection ratio increases
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the intercooling effect during compression process. Decreasing the discharge temperature of the compressor at high-ambient temperature helps enhance the reliability of the system. Therefore, in the case of high-ambient
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temperatures, use of the injection cycle in the multi-split VRF system is recommended to decrease the
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discharge temperature of the compressor and enhance the system’s reliability.
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5. Conclusions
Different operating characteristics of the multi-split VRF system with the bypass cycle and injection cycle are compared in this study based on the numerical simulation. The comparison results are summarized as follows: 1. The sub-cooling degree is influenced by the bypass and injection ratios because the overall heat transfer coefficient in a sub-cooler is determined by the bypass and injection ratios. In the bypass cycle, the degree of superheating at compressor suction is affected by the bypass ratio because of the mixing of the bypassed refrigerant with that from the evaporators. In the injection cycle, however, the degree of superheating at compressor suction is constant despite the exhibited variation in the injection ratio. In addition, the mass flow
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ACCEPTED MANUSCRIPT rate in the evaporator decreases as the bypass ratio increases. However, it is almost constant as the injection ratio increased. 2. The discharge temperature of the compressor decreases when the bypass and injection ratios increase. This is because the suction temperature decreased as the bypass ratio increases in the bypass cycle, and the intercooling effect is expected in a compression chamber during the compression process in the injection cycle.
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3. Both the bypass cycle and injection cycle yield improvements of the cooling capacities by 3.22 % and
13.43 %, respectively. In the bypass cycle, the cooling capacity decreases beyond the optimum bypass ratio owing to a decrease of the mass flow rate in the evaporators in spite of the increase of the sub-cooling degree at the inlet of the EEV. In the injection cycle, the cooling capacity is enhanced
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as the injection ratio increases because of the constant mass flow rate in the evaporators and the increase of the sub-cooling degree at inlet of the EEV.
4. The application of the injection cycle as the sub-cooling method reduces the input power because the compressor is operated at a lower rotational speed compared to the bypass cycle under the same capacity. In
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addition, the injection cycle is more effective in reducing the discharge temperature at high-ambient
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temperature conditions.
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ACKNOWLEDGMENTS
This work was supported by “Human Resources Program in Energy Technology” of the Korea Institute of
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Energy Technology Evaluation and Planning (KETEP), granted financial resource from the Ministry of Trade,
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Industry & Energy, Republic of Korea. (No. 20184010201660)
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Fig. 1 Schematic and P–h diagram for the bypass cycle and injection cycle.
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(a) Bypass cycle and (b) injection cycle
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Fig. 2 P-h diagram of injection cycle with a scroll compressor
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Fig. 3 Flow chart of the simulation logic
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Fig. 4 Experimental apparatus for the multi-split VRF systems
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Fig. 5 Validation results for the scroll compressor without the refrigerant injection.
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(a) Input power and (b) mass flow rate
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Fig. 6 Validation results of the simulation for the multi-split VRF system.
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(a) cooling mode and (b) heating mode
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(a) Sub-cooling degree at the inlet of EEVs
(b) Degree of superheating at exit of bleeding line and compressor suction
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Fig. 7 Sub-cooling degree and degree of superheating
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PT
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M
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Fig. 8 Mass flow rate. (a) total mass flow rate and (b) mass flow rate in the evaporator
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PT
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M
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Fig. 9 Discharge temperature
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(b) Input work
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M
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(a) Cooling capacity
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(b) Energy efficiency ratio (EER) Fig. 10 The operating characteristics of the system with variation of the bypass and injection ratio
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Fig. 11 Performance comparison for both cycles in the multi-split VRF system with the same capacity.
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(a) Input power/EER and (b) discharge temperature
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Pressure drop coefficient
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Heat transfer coefficient
Refrigerant side - Single phase: Gnielinski (1976) - Two phase : Condensation: Shah (1979) Evaporation: Kandliker (1990) Air side - Natural convection : Churchill and Chu. (1975) - Forced convection: Wang et al. (1999) Refrigerant side - Single phase: Churchill (1977) - Two phase : Condensation: Lockhart-Martinelli (1979) Evaporation: Gr𝑜̈ nnerud (1979) Air side - Forced convection: Wang et al. (1999)
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Table 1 Correlations of heat-transfer coefficients and pressure drop for the refrigerant and air sides
Table 2 Specifications of the multi-split VRF system Type 1 Compressor 1 ODU Fan
Outdoor unit
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Pipeline
Capacity : 22.4 kW × 1 unit Capacity : 2.5 kW × 9 units
Displacement volume : 48.0 𝑐𝑚3 /𝑟𝑒𝑣 Rotational speed : Max. 120 Hz Diameter of injection hole : 3.0 mm (located at 300° inward from end angle of wrap)
Louver fin-tube type
W1.75 m × H1.21 m (58 Columns, 2 Rows)
Louver fin-tube type
W1.23 m × H0.241 m (16 Columns, 3 Rows)
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Outdoor unit heat exchanger Indoor unit heat exchanger
Scroll compressor
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Compressor
Duct type
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Indoor unit
Specifications
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Components
Liquid line Copper tube Gas line
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Inner dia. : 11.074 mm Length : 100 m Inner dia. : 19.94 mm Length : 100 m
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Table 3 Uncertainty analysis of an experimental apparatus for the multi-split VRF system
Parameter
Unit
Accuracy of measurement devices
Uncertainty (%)
Suction pressure
kPa
±0.25 %
±0.32
Discharge pressure
kPa
±0.25 %
Air temperature
°C
±0.05 °C
Differential pressure transmitter
kPa
Input power
kW
Cooling capacity
kW
±0.1
±0.2 %
±0.44
3.43 ± 0.012
Heating
kW/kW
3.59 ± 0.017
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kW/kW
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±0.24 ±0.16
Cooling
35
±0.27
±0.07 %
kW
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Repeatability of COP for the system
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Heating capacity
±0.31
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Outdoor
100 %
Dry bulb Temp. (°C) 35.0
Wet bulb Temp. (°C) 23.9
Cooling
75 %
27.5
18.7
mode
50 %
20.0
13.9
25 %
18.3
Heating
Standard
8.3
mode
Low Temp.
-8.3
Dry bulb Temp. (°C)
Wet bulb Temp. (°C)
26.7
19.4
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Part load
Indoor
11.6 6.1
-9.4
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Mode
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Table 4 Experimental conditions
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21.1
15.6
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Operating setting Compressor Hz
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Table 5 Operating conditions for experiments and simulations Temperatures 65 650
IDU fan RPM
700
Superheating degree at outlet of Evap.
2 °C
Sub-cooling degree at outlet of Cond.
2 °C
Indoor
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PT
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M
ODU fan RPM
DB : 46.1 °C WB : 35.0 °C
Case 2
DB : 35.0 °C WB : 23.9 °C
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Outdoor
Case 1
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DB : 26.7°C WB : 19.4 °C
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ANSI/ASHRAE Standard 37-2005, 2005. Methods of Testing for Rating Electrically Driven Unitary Air-
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conditioning and Heat Pump Equipment. ASHRAE, Atlanta, GA, USA.
Cho, I., Seo, H., Kim, D., Kim, Y., 2016, Performance comparison between R410A and R32 multi-heat pumps with a sub-cooler vapor injection in the heating and cooling modes, Energy, Vol. 112, 179-187.
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Churchill, S.W. and Chu, H.H.S., 1975, Correlating equations for laminar and turbulent free convection from a vertical plate, Int. J. Heat Mass Transfer, Vol. 18(11), 1323-1329.
Jiang, H., Aute, V., Radermacher, R., 2006, CoilDesigner: a general-purpose simulation and design tool for
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air-to-refrigerant heat exchangers, International Journal of Refrigeration, Vol. 29 (4). 601-610.
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JSRAE, 2018, Compressors for Air Conditioning and Refrigeration, Technical Book Series, Japan Society of
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Refrigeration and Air Conditioning Engineers, 6-34.
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Kang, H., Joo, Y., Kim, Y., 2008, Effects of flash gas generation at the expansion device inlet on the dynamic
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characteristics of a refrigeration system, International Journal of Refrigeration. Vol.31(3), 396-403.
Kwon L., Hwang, Y., Radermacher, R., Kim, B., 2012, Field performance measurement of a VRF system with sub-cooler in educational building for the cooling season, Energy and Buildings, Vol. 49, 300-305.
Kwon C., Kim, M. S., Choi, Y., Kim, M. S., 2017, Performance evaluation of a vapor injection heat pump system for electric vehicles, International Journal of Refrigeration, Vol. 74, 138-150.
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ACCEPTED MANUSCRIPT Li, Z, Wang, B., Li, X., Shi, W., 2016, Simulation on effects of subcooler on cooling performance of multisplit variable refrigerant flow systems with different lengths of refrigerant pipeline, Energy and Buildings, Vol. 126, 301-309.
Liu, Y., Hung, C., Chang, Y., 2009, Mathematical model of bypass behaviors used in scroll compressor,
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Applied Thermal Engineering, Vol. 29 , 1058-1066.
Ragazzi F. and Pedersen C. O., 1996, Thermodynamic Optimization of Evaporators with Zeotropic
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Refrigerant Mixtures, ASHREA Transactions, 102, 367-372.
Shao, S., Xu, H., Tian, C., 2012, Dynamic simulation of multi-unit air conditioners based on two-phase fluid network model, Applied Thermal Engineering, Vol. 40, 378-388.
2017, A general simulation model for variable
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Sun, H., Ding, G., Hu, H., Ren, T., Xia, G., Wu, G.,
refrigerant flow multi-split air conditioning system based on graph theory, International Journal of
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Refrigeration, Vol. 82, 22-35.
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Tu Q., Zhang, L., Cai, W., Guo, X., Deng, C., Zhang, J., Wang, B., 2017, Effects of sub-cooler on cooling
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1152-1163.
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performance of variable refrigerant flow air conditioning system, Applied Thermal Engineering, Vol. 127,
Wang, X., Hwang, Y., Redermacher, R., 2009, Two-stage heat pump system with vapor-injection scroll compressor using R410A as a refrigerant, International Journal of Refrigeration, Vol. 32 (6), 1442-1451.
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Performance comparison between bypass cycle and injection cycle for sub-cooling methods in multi-split variable refrigerant flow (VRF) system in hot seasons Byungchae Min , Seokhoon Jang , Taemin Lee , Heunghee Bae , Cheoreon Moon , Gyungmin Choi PII: DOI: Reference:
S0140-7007(19)30308-1 https://doi.org/10.1016/j.ijrefrig.2019.07.010 JIJR 4465
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
23 October 2018 26 May 2019 18 July 2019
Please cite this article as: Byungchae Min , Seokhoon Jang , Taemin Lee , Heunghee Bae , Cheoreon Moon , Gyungmin Choi , Performance comparison between bypass cycle and injection cycle for sub-cooling methods in multi-split variable refrigerant flow (VRF) system in hot seasons, International Journal of Refrigeration (2019), doi: https://doi.org/10.1016/j.ijrefrig.2019.07.010
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 Highlights
The bypass cycle and injection cycle in a multi-split VRF system help increase the sub-cooling degree at inlet of EEV.
Energy efficiency ratio (EER) of the multi-split VRF system increases with application of the bypass
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cycle and injection cycle. Performance of the VRF system with the injection cycle is more efficient under the same cooling
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capacity.
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Performance comparison between bypass cycle and injection cycle for sub-cooling methods in multi-split variable refrigerant flow (VRF)
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system in hot seasons
Byungchae Mina, Seokhoon Jangb, Taemin Leea, Heunghee Baeb, Cheoreon Moonb, Gyungmin Choia* a
School of Mechanical Engineering, Pusan National University, Busan 46283,
b
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Republic of Korea
SAC Research/Engineering Division, H&A Solution company, LG Electronics, 84,
*
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Wanam-ro, Seongsan-gu, Changwon-si, Gyeongsangnam-do,51554, Republic of Korea
Corresponding Author: Professor Gyungmin Choi, School of Mechanical
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Engineering, Pusan National University, Busan 46283, Republic of Korea Telephone: +82-51-510-2476, Fax: +82-51-512-5236
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Email: [email protected]
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Abstract
In this study, the performance of a multi-split VRF system using the bypass cycle and injection cycle is
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evaluated using the numerical simulation as a possible sub-cooling method to prevent flash gas generation in liquid pipelines. The simulation for the multi-split VRF system is developed by considering the
applications of the bypass cycle and injection cycle, and is validated with experimental data. The bypass cycle and injection cycle in the multi-split VRF system yield improvements in their cooling capacities of the order of 3.22 % and 13.43 %, respectively, and energy efficient ratio (EER) of the order of 1.98 % and 1.72 %, respectively. The input power of the injection cycle is reduced by up to
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ACCEPTED MANUSCRIPT 4.45 % when the performance of the multi-split VRF systems with bypass cycle and injection cycle is compared under the same cooling capacity conditions.
Keywords: Variable refrigerant flow system, Sub-cooling method, Bypass cycle, Injection cycle, Numerical
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simulation
NOMENCLATURE : Area (m2)
C
: Flow coefficient (-)
COP
: Coefficient of Performance (-)
h
: Enthalpy (kJ kg-1 )
HT
: Heat transfer
k
: Ratio of specific heat (-)
m, M
: Mass (kg)
ṁ
: Mass flow rate (kg s-1)
P
: Pressure (kPa)
Q
: Capacity (kW) and heat transfer rate (kW)
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A
Rbyp
: Bypass ratio (-)
rcr
: Critical pressure ratio (-)
Rinj
: Injection ratio (-)
T
: Temperature (K)
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: Volume (m3)
v
: Specific volume (m3 kg-1)
W
: Input power (kW)
: Efficiency (%)
θ
: Angle (degree)
ρ
: Density (kg m-3)
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η
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Greek letter
Subscript : average
byp
: bypass
cal
: calculation
comp
: compressor
cond
: condenser
d
: downstream
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M
avg
dis
: discharge
EEV
: electronic expansion valve
eva
: evaporator
g
: gas-phase
4
ACCEPTED MANUSCRIPT : internal heat exchanger, sub-cooler
in
: inlet
ind
: indicated
inj
: injection
init
: initial
l
: liquid-phase
m
: mass
mech
: mechanical
motor
: motor
out
: outlet
pipe
: pipeline
sc
: sub-cooling
set
: setting
sh
: super-heating
suc
: suction
u
: upstream
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vi
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IHX
: vapor-injection
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ACCEPTED MANUSCRIPT Introduction The multi-split variable refrigerant flow (VRF) system has been used to serve cooling or heating units in the middle or large-sized buildings, such as residential and commercial buildings (ex. apartments, hotels, theaters, gyms, etc.), because this system can adjust the cooling or heating capacity for each separated zone according to the cooling or heating loads using an inverter scroll compressor and electronic expansion valves (EEV). In
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addition, the multi-split VRF system consists of at least one outdoor unit and several indoor units, which have different capacities and types, thereby reducing the installation space of outdoor units at the ground or rooftop of the buildings. However, pipelines between outdoor and indoor units can be lengthened by at most 150 m among existing VRF system products (Afify, 2008). Because the multi-split VRF system is used in the middle
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or large-sized buildings with long pipelines, the system should be designed by considering the pressure drop in the pipelines between the outdoor and indoor units since the pressure drop causes performance degradation of the overall system owing to the increased energy consumption required to transfer the working fluid and the flash gas generation in the liquid pipelines. Specifically, the flash gas generation in the liquid pipeline leads to
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noise in EEVs, decrease of the mass flow rate through the EEV, and yields lower system efficiency and reliability. Kang et al. evaluated the influence of flash gas generation in a refrigeration system. As the result of
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their study, the mass flow rate through an EEV decreased due to the increase of the void fraction with the increase of flash gas ratio. Hence, the cooling capacity and input power decreased with the increase of flash
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gas ratio because of reduced mass flow rate, and the coefficient of performance (COP) of the refrigeration system was degraded significantly since a decrease of the cooling capacity was larger than that of the input
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power with an increase of flash gas ratio (Kang et al., 2012). Therefore, in the multi-split VRF system, a sub-
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cooler heat exchanger is typically used to prevent flash gas generation in the liquid pipeline. In a multi-split VRF system, two types of refrigeration cycles that use the sub-cooler to prevent flash gas generation exist. The first one is a bypass cycle, and the other is an injection cycle. Fig. 1 shows a schematic of the bypass cycle and injection cycle. In the bypass cycle, the bled refrigerant is expanded at a low temperature, which is similar to the response of the evaporating temperature in the system. The low-temperature refrigerant in the bypass stream is heated by the heat exchange with a high temperature in the main stream, and the bled refrigerant is finally mixed with the refrigerant flowing out from the evaporator in the accumulator. In the injection cycle, the bled refrigerant
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ACCEPTED MANUSCRIPT is expanded to a higher pressure than the bypass pressure. The bled refrigerant is finally injected into a compression chamber in a scroll compressor. In both cycles, a larger sub-cooling degree of the system can be elicited with the sub-cooler compared to the system without the sub-cooler, thereby avoiding flash gas generation and increasing the cooling capacity. Moreover, when the bled refrigerant is in two-phase state and temperature of the bled refrigerant is not higher than that at the suction port of a compressor, the discharge
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temperature of a compressor can decrease due to a decrease of suction temperature of a compressor in the bypass cycle. In addition, the discharge temperature can decrease by an intercooling effect of the refrigerant injection in the injection cycle.
There are some studies that investigated the performance of a multi-split VRF system by using a sub-cooler
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heat exchanger with the bypass cycle and injection cycle in the literature. Kwon et al. investigated the effect of the sub-cooler heat exchanger with a bypass cycle on the performance of a multi-split VRF system, based on field performance tests in educational offices during the cooling season. In their study, it was revealed that a multi-split VRF system with a sub-cooler heat exchanger improved the cooling performance factor (CPF) by
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approximately 8.5 % compared to a system without a sub-cooler heat exchanger (Kwon et al., 2012). Li et al. analyzed the effect of a sub-cooler heat exchange in a multi-split VRF system with the bypass cycle by
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varying three factors, namely, the length of the pipeline, the capacity of the sub-cooler heat exchanger, and the bypass refrigerant mass flow rate ratio. Their simulation results indicated that the COP of a multi-split VRF
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system can be improved with a long pipeline by controlling the appropriate bypass mass flow rate ratio (Li et al., 2016). Tu et al. carried out experiments for the analyses of the effects of the sub-cooler on the cooling
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capacity, energy efficiency ratio, degree of sub-cooling, and degree of discharge superheating in a four-indoor
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VRF system with the bypass cycle. In accordance to their experimental results, they found that the reasonable adjustment of opening of the sub-cooler EEV can improve the performance of a multi-split VRF system by adopting the bypass cycle. The maximum energy efficiency ratio of the system with a sub-cooler was improved by up to 3.3 %, 3.1 %, and 3.0 %, when one, two, and four indoor units were operated on the thermo-on mode, respectively (Tu et al. 2017) . Wang et al. adopted an internal heat exchanger (sub-cooler) in a single-unit heat pump system with an injection cycle to enhance the performance of the system in hot and cold climates. Vapor injection in a scroll compressor improved the performance of the heat pump system because the heat exchange between the main and injected streams in the sub-cooler reduced the refrigerant
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ACCEPTED MANUSCRIPT enthalpy at the inlet of the evaporator with an increase of the degree of sub-cooling (Wang et al. 2009). Cho et al. measured and compared the cooling and heating capacities of R410A and R32 multi-heat pump systems with a vapor injection. The cooling and heating capacity of the system with R410A and R32 were enhanced by vapor injection owing to an increase of the sub-cooling degree at the inlet of the EEV using a sub-cooler (Cho et al., 2016).
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In the studies mentioned previously, it was found that use of a sub-cooler can increase the degree of subcooling for both bypass cycle and injection cycle. Therefore, flash gas generation in the liquid pipeline can be prevented by the sub-cooler. The bypass cycle and injection cycle can be applied to the same multi-split VRF system with a sub-cooler by using solenoid valves to switch the bled stream path into the accumulator or into
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the compression chamber. In addition, it can be found that the multi-split VRF system using the bypass cycle and injection cycle has different operating characteristics in the literature. However, the study on performance comparison between both cycles in the multi-split VRF system based on the numerical simulation cannot be found in the literature.
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The objective of this study is to compare the operating characteristics of both multi-split VRF systems using the bypass cycle and injection cycle based on the numerical simulation. The bypass and injection ratios are
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important factors to determine the performance of a multi-split VRF system. Therefore, the operating characteristics of both systems are compared by varying the bypass and injection ratios in the same
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configuration of the multi-split VRF system and by using the same operating conditions. Moreover, the improvement potential of two cycles in the multi-split VRF system is compared with optimum bypass ratio
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and injection ratio under same cooling capacity.
2. Mathematical models 2.1 Scroll compressor To simulate the performance of a compressor, three models were considered, such as the semi-empirical model (Sun et al., 2017, Zhu et al., 2013, Shao et al., 2012), map-based model (10-efficient ARI compressor map), and the geometry-based model (Kwon et al., 2017).The semi-empirical and map-based models have been used in most previous studies because these models are simpler in numerical simulations with experimental data to reduce the simulation time. However, these models are not suitable for the simulations of
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ACCEPTED MANUSCRIPT heat pumps equipped with a refrigerant-injection scroll compressor. This is because the refrigerant injection into a compression chamber in a scroll compressor occurs continually during the compression process, as shown in Fig. 2. This depends on many parameters, including the injection pressure, suction pressure, position of an injection hole, sub-cooler type and capacity, and others. Hence, Kwon et al. adopted the geometry-based thermodynamic model in the simulation of a vapor injection heat pump for electric vehicles in order to
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consider the vapor injection process in a scroll compressor. In the bypass cycle, the degree of superheating at compressor suction, which significantly affects the input power and discharged mass flow rate of a scroll compressor, is changed with the variation in the bypass mass flow rate ratio. Moreover, in the injection cycle, the input power and discharged mass flow rate of the scroll
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compressor are significantly affected by the injection process. In this study, therefore, the geometry-based thermodynamic model was applied to the simulation of the multi-split VRF system with a sub-cooler in order to take into account the refrigerant injection process and degree of superheating at compressor suction. The governing equations for the numerical model of a scroll compressor include changes in mass, enthalpy,
=
𝑑𝑚𝑖𝑛 𝑑𝜃
−
𝑑𝑚𝑜𝑢𝑡 𝑑𝜃
𝑑ℎ
+
𝑑𝑚𝑖𝑛𝑗 𝑑𝜃
(1)
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𝑑𝑚 𝑑𝜃
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pressure, and temperature in a pocket, as respectively indicated by Eqs. (1)–(2) (JSRAE, 2018).
𝑑𝑃
(2)
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M 𝑑𝑡 = 𝑚̇𝑖𝑛 (ℎ𝑖𝑛 − ℎ) − 𝑚̇𝑜𝑢𝑡 (ℎ𝑜𝑢𝑡 − ℎ) + 𝑚̇𝑖𝑛𝑗 (ℎ𝑖𝑛𝑗 − ℎ) + 𝑄 + 𝑉 𝑑𝑡
where Q is heat transfer rate between the wrap wall and the refrigerant in a pocket. The pressure and
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temperature in a control volume are calculated using the density and enthalpy obtained from Eqs. (1)-(2). REFPROP 9.0 was used to calculate the properties of the working fluids. In this study, the governing equations for the numerical model of the scroll compressor are formulated based on the following assumptions: a) Leakages are ignored between pockets b) Heat transfer and oil effects in a scroll compressor are ignored. Therefore, Q in Eq. (2) is zero kW c) Properties in a volume are homogeneous
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ACCEPTED MANUSCRIPT d) The gravity and kinetic energy are ignored Mass flow rates in a scroll compressor, such as suction, injection, and discharge flows, are predicted by Eqs. (3)–(5) (JSRAE, 2018)..
2/𝑘
2𝑘 𝑃 𝑚̇ = 𝐶𝐴√((𝑘−1) 𝑃𝑢 𝜌𝑢 ) {( 𝑃𝑑 )
−
}
(3)
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𝑢
𝑘+1
𝑃 𝑘 ( 𝑃𝑑 ) 𝑢
𝑘
𝑟𝑐𝑟 =
2 𝑘−1 (𝑘+1)
when
𝑃𝑑 𝑃𝑢
(4)
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< 𝑟𝑐𝑟 , then 2
𝑚̇ =
2𝑘 2 𝑘−1 𝐶𝐴√𝑘+1 𝑃𝑢 𝜌𝑢 {(𝑘+1) }
(5)
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When the refrigerant injection flow is two-phase flow (𝑥𝑖𝑛𝑗 < 1), the upstream density (𝜌𝑢 ) is calculated by
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Eq. (6) (JSRAE, 2018).
(6)
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𝜌𝑢 = 𝜌𝑙 + 𝑥(𝜌𝑔 − 𝜌𝑙 )
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Eq. (6) is a homogeneous model for two-phase flow. The injection mass flow rate is predicted by the sum of the injection mass flow rate at every crank angle. Furthermore, the bypass and injection ratios are defined as
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follows:
𝑅𝑏𝑦𝑝 =
𝑅𝑖𝑛𝑗 =
𝑚̇𝑏𝑦𝑝
(7)
𝑚̇𝑑𝑖𝑠
𝑚̇𝑖𝑛𝑗
(8)
𝑚̇𝑑𝑖𝑠
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ACCEPTED MANUSCRIPT The bypass ratio is the ratio of the discharged mass flow rate to the mass flow rate bypassed into the bleeding line. The injection ratio is the ratio of the discharged mass flow rate to the mass flow rate injected into a compression chamber. In the bypass cycle, summation of the mass flow rates bypassed and flowing in the evaporator is same with the suction mass flow rate which is same with the discharged mass flow rate, while in the injection cycle, the discharged mass flow rate is equal to summation of the suction and injection mass flow
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rates. The indicated work of a scroll compressor can be predicted by the variation in pressures and volumes of the pockets from the suction to the discharge processes, as indicated by Eq. (9).
𝑊𝑖𝑛𝑑 = ∮ 𝑃 𝑑𝑉
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(9)
Finally, the compressor input work is calculated using Eq. (10), including the motor and mechanical efficiencies. The motor, mechanical, and volumetric efficiencies are obtained from compressor calorimeter
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tests with variation in the operating conditions such as suction and discharge pressures, and rotational speed.
(10)
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𝑊𝑐𝑜𝑚𝑝 = 𝑊𝑖𝑛𝑑 /𝜂𝑚𝑜𝑡𝑜𝑟 𝜂𝑚𝑒𝑐ℎ
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The data of these efficiencies are formulated by the linear regression.
The injection flow area is calculated using an equation derived by Liu et al. in relation to the orbiting angle
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the involute.
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(Liu et al., 2009). The injection hole was located at 300° in the inward direction from the suction port along
2.2 Heat exchangers 2.2.1 Condenser and evaporator The multi-split VRF system adopted a louver fin-tube heat exchanger in an outdoor unit as a condenser when operated in the cooling mode. In this study, the tube-by-tube method is applied to the heat exchanger
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ACCEPTED MANUSCRIPT simulation to calculate the heat transfer between the refrigerant and air, and the pressure drop of the refrigerant and air sides in one tube defined as a control volume with the effectiveness–NTU method.
2.2.2 Sub-cooler A tube-in-tube heat exchanger type, which consists of one outer tube and seven inner tubes, is used as the sub-
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cooler. The flow area of the outer tube is obtained by the equivalent diameter to calculate the flow velocity, Reynolds number, and the heat transfer coefficient. The heat transfer in the sub-cooler is calculated by adopting the effectiveness-NTU method.
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2.3 Pipelines
The liquid pipeline and the gas pipeline are connected between the outdoor and indoor units. In the multi-split VRF system, the system performance is affected by the pressure drop and heat transfer in these pipelines. Hence, in this study, the pressure drop and the heat transfer in the pipelines are modeled to consider the effects
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of these factors on the overall system performance. The correlations used for the model of the heat exchangers
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and the pipelines are summarized in Table 1.
2.4 Electronic expansion valve (EEV)
= 𝐶𝑣 𝐴√𝜌
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𝑚̇
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The mass flow rate flowing through an EEV can be predicted by the orifice equation,
,𝑖𝑛
where A is the flow area,
𝑃
𝑃
(11)
is the pressure difference between the inlet and outlet of the EEV, 𝜌
,𝑖
is the
refrigerant density at the inlet, and 𝐶𝑣 is the flow coefficient which is obtained from the literature (Zhu et al., 2013). 2.5 Multi-split VRF system equipped with a sub-cooler In the simulation of a multi-split VRF system, the basic laws of mass, energy, and momentum conservation of the refrigerant properties in the entire system should be satisfied, as indicated by Eqs. (12) - (14).
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𝑚̇𝑐𝑜𝑚𝑝 − ∑ 𝑚̇
− 𝑚̇𝑖𝑛𝑗 = 0
(12)
𝑊𝑖𝑛,𝑐𝑜𝑚𝑝 + ∑𝑛𝑖=1 𝑄𝑒𝑣𝑎,𝑛 - 𝑄𝑐𝑜𝑛𝑑 = 0
(13)
𝑃𝑑𝑖𝑠 − 𝑃𝑠𝑢𝑐 =
(14)
+ 𝑃𝑝𝑖𝑝𝑒
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𝑃
Fig. 3 shows the flow chart of the simulation logic for the multi-split VRF system which can satisfy the governing equations of the entire system. As shown in Fig. 3, the simulation logic consists of five sub-loops
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for convergence based on the set criterion with the use of an iterative scheme. In the first loop, mass flow rate from the compressor and discharge enthalpy are predicted by the compressor model. The discharge enthalpy is used as the enthalpy at the inlet of the condenser with assuming that the heat transfer and pressure drop in the pipe between the compressor and condenser are neglected. Subsequently, the enthalpy at the outlet of the
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condenser is calculated by the condenser model. In order to satisfy the preset sub-cooling degree at the outlet of the condenser, the calculations of the compressor and condenser models are repeated by adjusting the
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condensing pressure. The second loop is for the convergence of the degree of superheating at exit of injection line, when the degree of superheating at exit of injection line is given as the input data. In this loop, the
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pressure drop and heat transfer in the liquid pipeline were predicted by the pipeline model. In the third loop,
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the mass flow rate was calculated through an EEV in each indoor unit. When the mass balance between the summation of the mass flow rate for all EEVs and the mass flow rates in the evaporators is not satisfied, the
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evaporating pressure is adjusted to match the summation of the mass flow rate for all EEVs with that in the evaporators. At the fourth loop, the cooling capacity is predicted based on the evaporator model using the inlet conditions of the evaporators obtained at the previous steps and assuming an evaporating pressure. If the superheating degree at the outlet of the evaporators is not matched with the set value, the opening areas of the EEVs are adjusted according to the current superheating degree. At the last loop, the pressure drop and heat transfer in the gas pipeline are calculated to obtain the degree of superheating at compressor suction. If the
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ACCEPTED MANUSCRIPT calculated suction enthalpy of a compressor is not the same as the initially assumed value, the assumed suction enthalpy is substituted by the calculated value. The iterative calculation is repeated until the relative and absolute errors are less than their acceptable values. The relative and absolute errors are defined in accordance to Eqs. (15)–(18).
𝑠𝑐,𝑎𝑣𝑔
−
𝑠𝑐,𝑠𝑒𝑡 |
0.1
𝑒𝑠𝑢𝑝,𝑒𝑣𝑎 = |
𝑠ℎ,𝑒𝑣𝑔
−
𝑠ℎ,𝑠𝑒𝑡 |
0.1
𝑒𝑠𝑢𝑝,
= |
𝑠ℎ,𝑖𝑛𝑗
−
𝑠ℎ,𝑖𝑛𝑗,𝑠𝑒𝑡 |
0.1
𝑒𝑚
=
| ∑ 𝑚̇
−𝑚̇ 𝑣 |
(15)
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𝑒𝑠𝑢𝑏,𝑐𝑜𝑛𝑑 = |
(16) (17)
0.1%
(18)
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𝑚̇ 𝑣
3. Experimental apparatus
The specifications for the object system are summarized in Table 2. The rated cooling capacity of the object
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VRF system in this study is 22.4 kW, which comprises one outdoor unit and nine indoor units. Fig. 4 shows the schematic of the experimental apparatus for the multi-split VRF system. The outdoor unit consists of a
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scroll compressor with the displacement volume of 48.0 cc 𝑟𝑒𝑣 −1, a fin-and-tube heat exchanger of the louver-fin type, sub-cooler with the tube-in-tube type, a receiver, an accumulator, EEVs, and a 4-way valve. In
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addition, solenoid valves are used in order to switch the injection cycle mode or the bypass cycle mode, or
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close the paths for the injection line and the bypass line. In order to store the remainder of refrigerant, a receiver is installed at outlet of the condenser in the outdoor unit.
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The indoor unit consists of a heat exchanger of the louver-fin type and an EEV to control the mass flow rate with respect to the degree of superheating at exit of the evaporator. Cooling or heating capacities of the multi-split VRF system are measured in the environmental chambers which can maintain the ambient temperature and humidity, and are composed with one outdoor chamber and two indoor chambers as shown in Fig. 4. The air flow rates of the indoor units are measured by a differential pressure transmitter across the standard nozzle. Power consumption of the outdoor unit and indoor units is measured by a power meter. The suction and discharge pressures are measured by a pressure transducer. Dry
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ACCEPTED MANUSCRIPT bulb and wet bulb temperatures are measured by RTD sensors. The uncertainties for the measurement devices are shown in Table 3. Measurement of the cooling and heating capacities for an indoor unit is repeated six times with three samples of the indoor unit. The repeatability is also shown in Table 3. The experiment on the performance of the multi-split VRF system is carried out in accordance with the AHRI Standard 1230. When the cooling performance was measured, the refrigerant injection technique was not uses
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in the VRF system, but it was used when the experiment on the heating performance was conducted. The experimental conditions according to the AHRI Standard 1230 are summarized in Table 4.
4. Results and discussion
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4.1 Validation of the simulation
The simulation results are compared with experimental data in order to confirm the reliability of the simulation developed in this study. First, the simulation of the scroll compressor is validated based on the compressor calorimeter test. However, in the calorimeter test, details on the injection process could not be
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measured. Therefore, the calculated mass flow rate and input work for the scroll compressor without the refrigerant injection are compared with the calorimeter test data. Figs. 5 (a) and (b) show the validation results
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for the scroll compressor without the refrigerant injection. As shown in Fig. 5, the simulation results for the input work and the mass flow rate of the scroll compressor without the refrigerant injection show good
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agreement with the results of the calorimeter test, and are within 15 % and 5 %, respectively. The simulation results for the multi-split VRF system in the cooling mode are compared to the experimental
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data obtained without the refrigerant injection. Fig. 6 (a) shows the validation results for the simulation of the
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multi-split VRF system in the cooling mode with variation in the part load. As shown in Fig. 5 (a), the simulation results for the cooling capacity and input power of the multi-split VRF system in cooling mode are consistent with experimental data within 15 %. Despite of validation of the system simulation in the cooling mode, it is also necessary to check the accuracy of the simulation for the multi-split VRF system to which the refrigerant injection is applied. The refrigerant injection in a scroll compressor was applied to the experiment for the multi-split VRF system in the heating mode. The validation results for the simulation of the system with an injection cycle in heating mode are indicated in Fig. 6 (b) and show good agreement with a maximum
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ACCEPTED MANUSCRIPT error of 10.3 %. As the results of the simulation, the injection ratios for two-operating conditions in heating mode are 11.6 % and 19.7 %, respectively.
4.2 Performance comparison of a multi-split VRF system using bypass cycle and injection cycle as subcooling methods
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In this study, the performance of a multi-split VRF system is compared according to the variations of the bypass and injection ratios when the sub-cooling methods are employed in the same configuration (length of pipes, capacity of sub-cooler, etc., as shown in Table 2), and the same operating conditions shown in Table 5 in the cooling season. The sub-cooling degree at the outlet of the condenser and the degree of superheating at
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the outlet of the evaporators are fixed as the input date in all cases of the simulation as shown in Table 5. In addition, the simulation was performed for two cases of ambient temperature according to the ASHRAE Standard 37 (ANSI/ASHREA Standard 37-2005, 2005).
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4.2.1 Sub-cooling degree at the inlet of EEV and degree of superheating
The sub-cooling degree at the inlet of the EEV and the degree of superheating at the exit of the bleeding line
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for the bypass cycle and injection cycle are considerably influenced by the bypass and injection ratios, as shown in Figs. 7 (a) and 7 (b) . The sub-cooling degree at the inlet of the EEV increases as the bypass and
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injection ratios increases, but these values become almost constant at the specific value of the bypass and injection ratios for both cycles in the two ambient temperature conditions, as shown in Fig. 7 (a). This is
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because the two-phase heat transfer area in the sub-cooler increases as the degree of superheating at the exit of
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the bleeding line decreases with increases of the bypass and injection ratios. The increase of the two-phase heat transfer area in the sub-cooler leads to the enhancement of the overall heat transfer coefficient according to the increase of the bypass and injection ratios. The heat transfer coefficient in the two-phase region is significantly higher than that in the single-phase. Therefore, the overall heat transfer coefficient is enhanced with the increase of two-phase heat transfer area, but it is not significantly changed after the degree of superheating at the exit of the bleeding line reaches 0 °C because the heat transfer in the sub-cooler occurs just in the two-phase region of the bleeding line without any more increase of two-phase heat transfer area, thereby leading to constant sub-cooling degree.
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ACCEPTED MANUSCRIPT In Fig. 7 (a), the sub-cooling degree of the bypass cycle is higher than that of the injection cycle because of lower temperature in the bleeding line and less mass flow rate in main line compared to the injection cycle. The bled refrigerant in the bypass line is expanded near the evaporating pressure. However, the injection pressure is higher than the evaporating pressure to inject the refrigerant into the compression chamber during the compression process. Hence, in the bypass cycle, the temperature difference in the sub-cooler is larger
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than that in the injection cycle. In addition, the mass flow rates in the main and bleeding lines in case of the bypass cycle are less than those in case of the injection cycle as shown in Fig. 8. Therefore, the sub-cooling degree at inlet of EEV and the degree of superheating at exit of bleeding line in the bypass cycle are higher compared to the injection cycle as shown in Fig. 7.
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As shown in Fig. 7 (a), in case of the baseline with the bypass and injection ratios of 0 %, the sub-cooling degree at the inlet of EEV is lower than the setting value at exit of the condenser with 2.0 °C owing to heat gain from the ambient and pressure drop in the liquid pipeline. In general, the heat gain and pressure drop in the liquid pipeline causes the flash gas generation at EEV with low sub-cooling degree. However, in both
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cycles, despite the heat gain and pressure drop, the sub-cooling degree at the inlet of EEV is higher than that at the exit of the condenser. Therefore, it is possible to prevent flash gas generation in the liquid pipeline with
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appropriate bypass and injection ratios using the sub-cooler in the both cycles.
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As shown in Fig. 7 (b), the bypass ratio shows also an important influence on the degree of superheating at compressor suction (point 1 in Fig. 1) for the bypass cycle, but the degree of superheating at compressor
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suction is not influenced by the variation in the injection ratio. In the bypass cycle, the bled refrigerant is mixed with refrigerant flowing out from the evaporators in the accumulator after the completion of the heat
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exchange with the refrigerant in the main stream. Hence, the degree of superheating at compressor suction is affected by the degree of superheating at the exit of the bypass line. However, in the injection cycle, the degree of superheating at compressor suction is constant despite the variation in the injection ratio because the bled refrigerant is injected directly into a compression chamber. The constant degree of superheating at compressor suction is because of using the liquid receiver tank in the multi-split VRF system as shown in Fig. 4. This receiver tank is designed to hold the liquid refrigerant in order to vary the refrigerant flow rate in the overall system. In the experiment on the multi-split VRF systems, the refrigerant flow rate in the evaporator
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ACCEPTED MANUSCRIPT can be maintained by the store of the refrigerant in the receiver tank despite of the increase of injection ratio. In the simulation, it is assumed that the refrigerant charge is sufficient to supply it into the evaporators. In the bypass cycle, the receiver tank is also used. However, the suction mass flow rate is determined by the compressor displacement volume, rotational speed, and suction density. Therefore, the mass flow rate in the evaporators decreases with the increase of the bypass ratio.
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Fig. 8 shows the total mass flow rate discharged from the compressor and the mass flow rate in the evaporators. In the case of the bypass cycle, the total mass flow rate is not changed significantly at the beginning of the bypass, but increases slightly with additional increases of the bypass ratio because of the decrease of the degree of superheating at compressor suction. The mass flow rate in the evaporators decreases
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continually with the increase in the bypass ratio because of the increase of the bypassed mass flow rate. In the case of the injection cycle, the total mass flow rate increases continually with the increase of injection ratio because the injected refrigerant is added to the suction mass flow rate. The evaporating pressure is not affected significantly by the injection ratio. Therefore, the variation in the suction density of the compressor is very
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small.
Fig. 9 shows the discharge temperature for the two cycles and the two ambient temperature conditions as a
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function of the bypass and injection ratios. The discharge temperature slightly increases at the beginning of the bypass and injection for both cycles at the two ambient temperature conditions, but the temperature starts
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to decrease as the bypass and injection ratios increase further. The reason for the decrease of the discharge temperatures is different for each cycle. In the bypass cycle, the discharge temperature is influenced by the
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suction temperature. However, in the injection cycle, it is influenced by the temperature of the injected
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refrigerant owing to the intercooling effect. Therefore, in the bypass cycle, the discharge temperature increases at the beginning of the bypass because of increase of the degree of superheating at compressor suction, but the discharge temperature decreases as the bypass ratio increases further owing to decrease of the degree of superheating at compressor suction as shown in Fig. 7 (b). In the injection cycle, the discharge temperature also increases at the beginning of the injection owing to high temperature of the injected refrigerant. However, the discharge temperature decreases as the injection ratio increases further because the intercooling effects in the compression chamber is enhanced owing to the lower temperature of the injected refrigerant and the increase of injection mass flow rate. When the ambient temperature is high as the case of 46.1 °C, the
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4.2.2 Cooling capacity and input power Fig. 10 (a) shows the cooling capacities for both cycles at two ambient temperature conditions. The variation trends for both cycles are different as the bypass and injection ratios increase. In the bypass cycle, the cooling capacity increases with the increase of the bypass ratio. However, it decreases from the specific bypass ratio in
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the two cases of different ambient temperatures. This is because the mass flow rate in the evaporators decreases continually despite the increase of the sub-cooling degree with the increase of the bypass ratio. In the injection cycle, the cooling capacity increases continually with the increase of the injection ratio for the two cases of ambient temperature since the mass flow rate in the evaporators are almost constant despite the
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increase of the injection ratio, while the sub-cooling degree increases as the injection ratio increases. The cooling capacity in the injection cycle increases rapidly after the two-phase injection starts because the density
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of the two-phase state is considerably larger than that of the vapor state. The input power is also influenced by the bypass and injection ratios as shown in Fig. 10 (b). In the case of the
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bypass cycle, the input power increases slightly with an increase of the bypass ratio because of the increase of the suction temperature of the compressor and the total mass flow rate, but it decreases as the bypass ratio
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increases further owing to the decrease of the suction temperature. In contrast, the input power in the injection
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cycle increases continually as the injection ratio increases because of the continuous increase of the total mass flow rate. The input power for the injection cycle also increases rapidly after the two-phase injection starts. The cooling capacities in both cycles decrease with the increase of the ambient temperature because the enthalpy at the inlet of evaporator increases owing to the high condensing saturated temperature. The input powers increase also in the high ambient temperature because of higher compression ratio. Fig. 10 (c) shows energy efficiency ratio (EER) as a function of the bypass and injection ratios. EER for the bypass cycle increases gradually with the bypass ratio, but the EER decreases from the specific bypass ratio because the cooling capacity in the two cases of different temperatures decreases despite of reduction of the
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injected refrigerant. The compressor consumes more compression work with larger specific volume of the refrigerant vapor. However, EER of the injection cycle slightly increases when the degree of superheating at the injection line is reached to 0 °C because of the decrease of specific volume for the injected refrigerant. However, EER of the injection cycle decreases again because the compression work increases sharply with
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starting of the two-phase injection. The highest EER for the bypass cycle in two ambient temperature conditions is attained at the bypass ratio of 15.0 % because the cooling capacities which are influenced by the mass flow rate are highest at this bypass ratio. The EER for the bypass cycle is enhanced up to 1.19 % and 1.98 %, respectively in two ambient temperatures compared to the baseline cycle with the bypass ratio of 0 %.
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In case of the injection cycle, the EER is highest when the degree of superheating of injection line is close to 0 °C, which corresponds to the injection ratios of 10.67 % and 16.01 %, respectively in two ambient
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temperatures. The EER for the injection cycle is enhanced up to 0.42 % and 1.72 % with those injection ratios compared to the baseline cycle with the injection ratio of 0 %. The EER of the bypass cycle is higher than that
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of the injection cycle. However, the cooling capacity for the injection cycle is considerably higher than that of
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the bypass cycle because of constant mass flow rate and the increase of enthalpy difference in evaporators.
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4.3 Performance comparison for the bypass cycle and injection cycle with the same cooling capacity In this section, the performance for both cycles is compared for the same cooling capacity condition at two different ambient temperature conditions. As shown in Fig. 10 (a), the cooling capacity of the injection cycle is higher than the maximum capacity of the bypass cycle. For the comparison between the bypass cycle and injection cycle with the same capacity, the cooling capacity of the injection cycle is matched to the maximum capacity of the bypass cycle because EER of the bypass cycle is higher than that of the injection cycle. The maximum capacities of the bypass cycle for 35.0 °C and 46.1 °C are reached with the bypass ratios of 15.0 %. In order to achieve those maximum capacities in the injection cycle the rotational speed of the compressor is
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ACCEPTED MANUSCRIPT adjusted. In addition, the degree of superheating at the exit of the injection line is set at 2.0 °C, thus allowing the adjustment of the injection pressure based on the iterative method in accordance to the simulation logic, as indicated in Fig. 3. Figs. 11 (a) and (b) show the comparison results for both cycles at the same cooling capacity condition for two ambient temperatures. The same capacities for the injection cycle are attained at the injection ratios of 10.71 %
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and 15.23 %, respectively. The input power of the injection cycle is reduced compared to that of the bypass cycle for 35.0 °C and 46.1 °C, which corresponds to 2.32 % and 4.45 %, respectively. This is because the rotational speed of the compressor could be reduced in order to attain the same cooling capacity as the maximum capacity achieved in the bypass cycle with the compressor rotational speed of 60 Hz and 57 Hz. As
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the input power is reduced in the injection cycle, the EER is also enhanced at the two ambient temperatures by up to 2.33 % and 5.08 %, respectively. In addition, the injection cycle is more effective in decreasing the discharge temperature at high-ambient temperatures, as shown in Fig. 11 (b). Specifically, the discharge temperature is reduced considerably in the injection cycle at the high-ambient temperature condition with the
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same cooling capacity. This is because that in the injection cycles, the same capacities for 35.0 °C and 46.1 °C are attained with the injection ratios of 10.7 % and 15.2 %, respectively. The higher injection ratio increases
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the intercooling effect during compression process. Decreasing the discharge temperature of the compressor at high-ambient temperature helps enhance the reliability of the system. Therefore, in the case of high-ambient
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temperatures, use of the injection cycle in the multi-split VRF system is recommended to decrease the
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discharge temperature of the compressor and enhance the system’s reliability.
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5. Conclusions
Different operating characteristics of the multi-split VRF system with the bypass cycle and injection cycle are compared in this study based on the numerical simulation. The comparison results are summarized as follows: 1. The sub-cooling degree is influenced by the bypass and injection ratios because the overall heat transfer coefficient in a sub-cooler is determined by the bypass and injection ratios. In the bypass cycle, the degree of superheating at compressor suction is affected by the bypass ratio because of the mixing of the bypassed refrigerant with that from the evaporators. In the injection cycle, however, the degree of superheating at compressor suction is constant despite the exhibited variation in the injection ratio. In addition, the mass flow
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3. Both the bypass cycle and injection cycle yield improvements of the cooling capacities by 3.22 % and
13.43 %, respectively. In the bypass cycle, the cooling capacity decreases beyond the optimum bypass ratio owing to a decrease of the mass flow rate in the evaporators in spite of the increase of the sub-cooling degree at the inlet of the EEV. In the injection cycle, the cooling capacity is enhanced
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as the injection ratio increases because of the constant mass flow rate in the evaporators and the increase of the sub-cooling degree at inlet of the EEV.
4. The application of the injection cycle as the sub-cooling method reduces the input power because the compressor is operated at a lower rotational speed compared to the bypass cycle under the same capacity. In
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addition, the injection cycle is more effective in reducing the discharge temperature at high-ambient
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temperature conditions.
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ACKNOWLEDGMENTS
This work was supported by “Human Resources Program in Energy Technology” of the Korea Institute of
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Energy Technology Evaluation and Planning (KETEP), granted financial resource from the Ministry of Trade,
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Industry & Energy, Republic of Korea. (No. 20184010201660)
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Fig. 1 Schematic and P–h diagram for the bypass cycle and injection cycle.
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(a) Bypass cycle and (b) injection cycle
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Fig. 2 P-h diagram of injection cycle with a scroll compressor
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Fig. 3 Flow chart of the simulation logic
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Fig. 4 Experimental apparatus for the multi-split VRF systems
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Fig. 5 Validation results for the scroll compressor without the refrigerant injection.
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(a) Input power and (b) mass flow rate
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Fig. 6 Validation results of the simulation for the multi-split VRF system.
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(a) cooling mode and (b) heating mode
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(a) Sub-cooling degree at the inlet of EEVs
(b) Degree of superheating at exit of bleeding line and compressor suction
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Fig. 7 Sub-cooling degree and degree of superheating
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Fig. 8 Mass flow rate. (a) total mass flow rate and (b) mass flow rate in the evaporator
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Fig. 9 Discharge temperature
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(b) Input work
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(a) Cooling capacity
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(b) Energy efficiency ratio (EER) Fig. 10 The operating characteristics of the system with variation of the bypass and injection ratio
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Fig. 11 Performance comparison for both cycles in the multi-split VRF system with the same capacity.
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(a) Input power/EER and (b) discharge temperature
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Pressure drop coefficient
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Heat transfer coefficient
Refrigerant side - Single phase: Gnielinski (1976) - Two phase : Condensation: Shah (1979) Evaporation: Kandliker (1990) Air side - Natural convection : Churchill and Chu. (1975) - Forced convection: Wang et al. (1999) Refrigerant side - Single phase: Churchill (1977) - Two phase : Condensation: Lockhart-Martinelli (1979) Evaporation: Gr𝑜̈ nnerud (1979) Air side - Forced convection: Wang et al. (1999)
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Table 1 Correlations of heat-transfer coefficients and pressure drop for the refrigerant and air sides
Table 2 Specifications of the multi-split VRF system Type 1 Compressor 1 ODU Fan
Outdoor unit
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Pipeline
Capacity : 22.4 kW × 1 unit Capacity : 2.5 kW × 9 units
Displacement volume : 48.0 𝑐𝑚3 /𝑟𝑒𝑣 Rotational speed : Max. 120 Hz Diameter of injection hole : 3.0 mm (located at 300° inward from end angle of wrap)
Louver fin-tube type
W1.75 m × H1.21 m (58 Columns, 2 Rows)
Louver fin-tube type
W1.23 m × H0.241 m (16 Columns, 3 Rows)
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Outdoor unit heat exchanger Indoor unit heat exchanger
Scroll compressor
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Compressor
Duct type
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Indoor unit
Specifications
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Components
Liquid line Copper tube Gas line
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Inner dia. : 11.074 mm Length : 100 m Inner dia. : 19.94 mm Length : 100 m
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Table 3 Uncertainty analysis of an experimental apparatus for the multi-split VRF system
Parameter
Unit
Accuracy of measurement devices
Uncertainty (%)
Suction pressure
kPa
±0.25 %
±0.32
Discharge pressure
kPa
±0.25 %
Air temperature
°C
±0.05 °C
Differential pressure transmitter
kPa
Input power
kW
Cooling capacity
kW
±0.1
±0.2 %
±0.44
3.43 ± 0.012
Heating
kW/kW
3.59 ± 0.017
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kW/kW
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±0.24 ±0.16
Cooling
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±0.27
±0.07 %
kW
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Repeatability of COP for the system
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Heating capacity
±0.31
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Outdoor
100 %
Dry bulb Temp. (°C) 35.0
Wet bulb Temp. (°C) 23.9
Cooling
75 %
27.5
18.7
mode
50 %
20.0
13.9
25 %
18.3
Heating
Standard
8.3
mode
Low Temp.
-8.3
Dry bulb Temp. (°C)
Wet bulb Temp. (°C)
26.7
19.4
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Part load
Indoor
11.6 6.1
-9.4
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Mode
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Table 4 Experimental conditions
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Operating setting Compressor Hz
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Table 5 Operating conditions for experiments and simulations Temperatures 65 650
IDU fan RPM
700
Superheating degree at outlet of Evap.
2 °C
Sub-cooling degree at outlet of Cond.
2 °C
Indoor
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ODU fan RPM
DB : 46.1 °C WB : 35.0 °C
Case 2
DB : 35.0 °C WB : 23.9 °C
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Outdoor
Case 1
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DB : 26.7°C WB : 19.4 °C
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