Investigation on the feasibility and performance of transcritical CO2 heat pump integrated with thermal energy storage for space heating

Accepted Manuscript Investigation on the feasibility and performance of transcritical CO2 heat pump integrated with thermal energy storage for space heating Zhihua Wang, Fenghao Wang, Zhenjun Ma, Wenye Lin, Haoshan Ren PII:

S0960-1481(18)31355-7

DOI:

https://doi.org/10.1016/j.renene.2018.11.035

Reference:

RENE 10798

To appear in:

Renewable Energy

Received Date: 20 April 2018 Revised Date:

6 October 2018

Accepted Date: 10 November 2018

Please cite this article as: Wang Z, Wang F, Ma Z, Lin W, Ren H, Investigation on the feasibility and performance of transcritical CO2 heat pump integrated with thermal energy storage for space heating, Renewable Energy (2018), doi: https://doi.org/10.1016/j.renene.2018.11.035. 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 1

Investigation on the feasibility and performance of transcritical CO2 heat pump

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integrated with thermal energy storage for space heating

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ZhihuaWanga*, FenghaoWanga*, Zhenjun Mab, Wenye Linb, Haoshan Renb

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School of Human Settlements and Civil Engineering, Xi’an Jiaotong University, Xi’an, Shaanxi, 710049, China

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Sustainable Buildings Research Centre (SBRC), University of Wollongong, Wollongong 2522, NSW, Australia

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Abstract

CO2 heat pumps have drawn a great deal of attention as an economic form of heating under low

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ambient temperature conditions. However, the system performance is not desirable and shows a lower

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COP due to the higher inlet water temperature at the gas cooler, which causes a higher refrigerant

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temperature at the exit of the gas cooler, leading to a large throttle loss when the refrigerant flow

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through the throttling device. To tackle this issue, a transcritical CO2 heat pump unit integrated with

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two thermal energy storage (TES) containing phase change materials (PCMs) is proposed in this paper.

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The objective of this work is to model and simulate the proposed system using TRNSYS based on a

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typical single family rural house in Beijing (typical cold climate conditions), China. The results showed

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that the heating capacity and energy consumption decreased by 21 and 24%, respectively, and the

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heating seasonal performance factor (HSPF) of the proposed system increased by 4% in comparison

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with the baseline system during the entire heating period. The simulation results demonstrated that TES

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is helpful to improve CO2 heat pump system performance and monthly energy saving ratio for space

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heating.

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Keywords: CO2 heat pump; TRNSYS; Simulation; Space heating; HSPF

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* Corresponding author. Tel:+86-13227006940; Fax:+86-29-83395100. E-mail address: [email protected] (Fenghao Wang); [email protected](Zhihua Wang).

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ACCEPTED MANUSCRIPT Nomenclature

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A

Heat transfer area, [m2];

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Cp

Specific heat of the heat transfer fluid, [kJ/(kg•K)];

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COPicycle

System COP without considering the effect of frosting;

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COPreddefrost

System COP considering the effect of frosting;

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COPcorr

System COP which was taken into account the effect of frosting;

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L

Length of the tube, [m];

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NTU

Number of transfer units;

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Pmax

Maximum power input, [kW];

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Qcorr

Correct system heating capacity which was taken into account the effect of frosting;

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Qmax

Maximum heating capacity, [kW];

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RT

Total thermal resistance, [K/W];

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Ri

Inner radius of the tube, [m];

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Ro

Outer radius of the tube, [m];

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S

Shape factor of the PCM, [m];

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Ta

Ambient temperature, [°C];

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Tin

Inlet water temperature in gas cooler, [°C];

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Tout

Outlet water temperature in gas cooler, [°C];

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U

Overall heat-transfer coefficient, [W/(m2•K)];

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h1

Inlet refrigerant enthalpy value of compressor, [kJ/K•kg];

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h2

Outlet refrigerant enthalpy value of compressor, [kJ/K•kg];

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h3

Inlet refrigerant enthalpy value of gas cooler, [kJ/K•kg];

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h4

Outlet refrigerant enthalpy value of gas cooler, [kJ/K•kg];

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Outlet refrigerant enthalpy of throttle vale at isoenthalpic throttling process, [kJ/K•kg];

h4’

Outlet refrigerant enthalpy of throttle vale at isentropic throttling process, [kJ/K•kg];

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hf

Heat transfer coefficient of the heat transfer fluid, [W/(m2•K)];

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m

Mass flow rate of heat transfer fluid, [kg/s];

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q

Heating capacity per mass flow rate, [W];

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w

Power input, [W];

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δ

Phase change fraction for a tube surrounded by a cylindrical volume of PCM

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ε

The heat exchanger effectiveness;

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µf

Dynamic viscosity of the heat transfer fluid, [kg/(m•s)];

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kw

Thermal conductivity of the tube wall [W/(m•K)];

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kPCM

Thermal conductivity of the PCM [W/(m•K)];

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Nu

Nusselt number;

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Re

Reynolds number;

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Pr

Prandtl Number;

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Abbreviations

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ASHP

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ASHPWH

Air-source heat pump water heater;

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COP

Coefficient of performance;

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DB

Dry-bulb temperature;

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WB

Web-blub temperature;

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HSPF

Heating seasonal performance factor;

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Air-source heat pump;

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1. Introduction In recent years, the problems of energy shortage and environmental pollution are among the two

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major challenges in sustainable development of global economy. The building energy consumption

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accounts for over 30% [1, 2] of the total energy consumption and it is growing at an annual rate of

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nearly 40% [3]. The building energy consumption at the rural regions accounts for approximately 37%

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of total building energy consumption. Particularly in the field of the heating energy consumption in

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northern rural building that it is responsible for over 60% of the total building energy consumption

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[4]. The heating demand in the rural regions is mostly produced by direct consumption of fossil fuels

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which results in environmental pollution and waste of energy. Hence it is essential to develop high

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performance heating equipment to reduce building energy consumption and mitigate environment

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pressure.

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In order to improve the current situation, the Chinese government has taken a series of financial

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incentives policies for promoting renewable energy technologies to facilitate the uptake of clean energy

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systems in the residential sector. The project of transform heating from coal to heating from electricity

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is one of them, i.e. taking air source heat pump (ASHP) to replace coal fired boiler for heating in rural

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areas [5]. Heat pumps, a device that use a small amount of external power to accomplish the work of

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transferring energy from the heat source to the heat sink, are proposed as a substitute for a baseline

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system, such as electric water heater, gas water heater, coal stove, etc. to produce domestic hot water

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and space heating [6]. It means that heat pump technology is promising mean of reducing the

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consumption of fossil energy resources and greenhouse gas emission [7, 8]. While ASHP, absorbed

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heat from outside air, is a type of heat pump, which is seen as particularly attractive option because of

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their flexible and convenient installation, compact in design relatively inexpensive in comparison with

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other types of heat pump such as ground source heat pump. Therefore, ASHPs have been widely used

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for preparing domestic hot water and space heating in China [9]. In rural regions, most heating systems used radiators as the terminal units. The supply/return water

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temperatures were 75/50ºC (Chinese Standard
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conditioning of civil buildings> GB 50736-2012). However, the outlet water temperature from the

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conventional ASHP unit was only 41ºC at the nominal design condition (-12DB/-14WB) (Chinese

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Standard < Low ambient temperature air source heat pump (water chilling) packages - Part 2: Heat

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pump (water chilling) packages for household and similar application > GB/T 25127.2-2010) that is

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lower than the supply water temperature. In order to satisfy the demand of indoor thermal comfort, it is

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usually to increase the numbers of radiators. Yang et al. [10] reported an application of an ASHP

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replaced the coal fired boiler for heating system in rural areas of Beijing, China. The result showed that

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the areas of the radiators were 19.6 m2 at the supply/return water temperature of 70/55ºC (the heat

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source was coal fired boiler) in the heating areas of 98.6 m2. When the heat source was changed to

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ASHP unit, the areas of radiators were 31.64 m2 that was 1.61 times of the original heat source at the

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supply/return water temperature of 45/45ºC. It increased the cost of investment and the difficulty of

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reconstruction.

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CO2, as a natural refrigerant, attracted much attention owing to its outstanding features with

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nontoxic, nonflammable low cost, moreover, it is characterized by high values of the thermal

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conductivity and heat capacity as well as a low value of its dynamic viscosity. Also, the high

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volumetric capacity of CO2 reduces component sizes. Since its low critical temperature (31.1°C), the

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reverse cycle behaves is different from the conventional one. The high pressure side is higher than the

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critical one and heat transfer of refrigerant does not involve two phase transformation but only gas 6

ACCEPTED MANUSCRIPT cooling, i.e. “transcritical cycle”. In addition, a transcritical CO2 heat pump unit can reduce the

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irreversible energy loss that leads to a higher performance because of the refrigerant temperature glide

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at heat rejection in gas cooler contributes to a very good temperature adaptation for heating a finite

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stream of water [11], resulting in a fairly large temperature lift in water without significant penalization

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in coefficient of performance (COP). So it achieves higher COP than that of the conventional

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refrigerants in the case of heat pump water heater. A host of publications in recent years have

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demonstrated that the refrigeration/heat pump system using CO2 as refrigerant is very competitive with

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conventional equipment for hot water heating, air-conditioning etc. regard to power consumption,

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compactness and cost when it operated in the transcritical region [12]. Nekså [13] pointed that the heat

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pump using CO2 as working fluid were gaining interesting in the technological and scientific. Saikawa

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et al. [14] evaluated the performance of an ASHPWH (air source heat pump water heater) system with

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various refrigerant in a relative value to the COP of the ideal cycle when tap water was heated from 17

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to 65ºC at the ambient temperature of 16ºC. It was found that, among the refrigerant of R134a, R22,

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R410A, R407C, R32 as Fluorocarbons, CO2, Propane, Isobutane and Propylene as natural refrigerants,

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CO2 achieved the highest COP for tap water heating. Cecchinato et al. [15] reported a theoretical

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comparison of R134a and CO2 in a tap water heat pump, it was found that CO2 is an interesting

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substitute for synthetic fluids because of its beneficial properties. Stene [16] studied a residential

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brine-to-water CO2 heat pump equipped with a unique counter-flow tripartite gas cooler for preheating

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of domestic hot water by theoretical analysis and experiments. It was found that the system COP

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dropped by approximately 15% when the inlet water temperature increased from 5 to 20°C at the

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supply/return temperatures of 35/30°C for space heating and the set-point temperature for domestic hot

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water of 60 °C. An experimental comparison of a prototype CO2 heat pump and commercially available

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R410A heat pump system in heating mode was conducted by Richter et al. [17]. It was found that the

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CO2 system operated with a slightly lower COP at testing condition, but its higher capacity at lower

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ambient temperatures reduced the need for less efficient supplementary heating capacity than R410A. From the aforementioned research, the CO2 heat pumps attracted increasing attention to develop

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renewable energy and energy savings in hot water heating since the temperature difference between the

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inlet and outlet water is relative higher. However, the system performance is not desirable and shows a

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lower COP when it is used for space heating. The reason for this phenomenon is that the higher inlet

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water temperature at the gas cooler caused a higher refrigerant temperature at the exit of the gas cooler,

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leading to a large throttle loss when the refrigerant flowing through the throttling device. As Lorentzen

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[18] pointed out that the system performance was decreased when the gas cooler side temperature

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increased. It is an inherent drawback of CO2 heat pump systems. Cycle performance improvements can

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be achieved by the methods such as multistage compression, expansion work recovery [19], throttling

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loss reduction by internal heat exchange [20] or ejector [21]. For space heating, where the CO2 heat

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pump unit gave off heat to a radiator heating system at supply/return water temperature of 75/50ºC, the

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throttle loss become more pronounced since higher return water temperatures would lead to operation

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at a lower COP.

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To best exploit CO2 heat pump unit potential of energy efficiency, it is mandatory to keep as low

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as possible the inlet water temperature at the gas cooler. Therefore, reducing the gas cooler water inlet

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temperature is one of the most promising methods to improve the CO2 heat pump system efficiency and

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reduce the throttling losses. Hence, a series of experimental and theoretical analyses for improving the

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efficiency of CO2 heat pump system by using sub-cooling technology that reduced the optimum heat

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rejection pressure and increasing the specific cooling capacity. Llopis et al. [22] published a CO2 8

ACCEPTED MANUSCRIPT transcritical refrigeration systems combined a dedicated mechanical sub-cooling cycle, the finding

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showed a maximum theoretical COP and cooling capacity improvement of up to 20 and 28.8%,

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respectively, over the conventional cycle at evaporating temperature from 5 to -30°C and environment

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from 20 to 35°C. Sarkar [23] presented a CO2 refrigeration cycle with thermoelectric sub-cooler

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observed that it not only improved the cooling capacity and COP, but also reduced the high-side

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pressure, compressor pressure ratio etc. It also found that improvements of COP up to 25.6% at studied

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range.

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In recent decades, thermal energy storage (TES) in the form of latent heat using PCMs has been

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widely used in building envelopes [24], solar facilities [25], heating, ventilating and air conditioning

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(HVAC) systems 26] because of its special thermal physical properties. It can store the surplus energy

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and release it when needed to improve the energy utilization efficiency. Many studies have been

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conducted to analyze the performance of PCM systems through various experimental and theoretical

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methods. Lu et al. [27], for instance, reported a solar water heating system coupled with the PCM floor.

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The result showed that the building using the heating system can save energy consumption of 5.87%

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when the indoor temperature was maintained at 20°C. Belmonte et al. [28] proposed a solar air-base

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system integrated with a fluidized bed energy storage unit containing PCMs. The results indicated that

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the system can successfully ensure part of the heating requirements for a single family house at mild

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winter conditions. Wang et al. [29] utilized a TES to absorb condenser heat and it acted as a low heat

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resource to regenerate solid desiccant for a frost-free ASHP system. Real et al. [30] exposed a HVAC

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system with two TES systems. The simulation results showed that the energy saving of 18.97% can be

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obtained.

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As mentioned above, using sub-cooling technology can improve the CO2 heat pump unit. TES is a 9

ACCEPTED MANUSCRIPT promising device for energy storage and transfer. Therefore, the advantage of combining of them, on

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the one hand, recovers the surplus energy of return water and decreases the refrigerant temperature at

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the exit of the gas cooler, where decreased the throttle loss and improved the system performance. On

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the other hand, the ambient temperature is relative low at night that deteriorates the heat pump system

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performance. In this situation, the heat pump system can be turned off and use the energy stored in TES

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for space heating which reducing tremendous energy consumption compared with the conventional

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heating space system. However, the studies related to conventional CO2 heat pump integrated with TES

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for spacing heating in cold rural regions are very limited. Therefore, the main objective of the present

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work is to critically analyze the heating performance of a CO2 heat pump system integrated with two

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TES systems, associated with high energy densities provided by the latent heat of the PCM, to meet the

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heating requirement of a rural single-family in Beijing, China. A transient system energy modeling

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software (TRNSYS 17.0), is used to evaluate and optimize the system performance, which has been

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verified in a number of studies such as in Refs [31, 32]. In the following sections, firstly, the

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performance for a theoretical cycle of CO2 heat pump system is analyzed. Secondly, the working

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principle of this kind of the CO2 heat pump heating system associated with TES for residential use is

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presented. Thirdly, a detailed description of the simulation model of the house, the proposed system,

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TES unit and control strategies in TRNSYS 17.0 are provided. Next, the simulation results of main

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output variables, such as variations in the indoor temperature, supply/return water temperature, system

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COP and HSPF (heating seasonal performance factor) etc., are reported. Finally, the representative

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conclusions are summarized. This study will be helpful to facilitate the development of high efficiency

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CO2 heat pump heating systems in cold regions.

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2． ．Methodology and Simulation.

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2.1 Performance analysis of a theoretical cycle of transcritical CO2 heat pump system Fig. 1 shows a typical one-stage transcritical CO2 vapor compression process in the

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temperature-enthalpy (t-h) diagram. It mainly consisted of a reciprocating compressor, a counter-flow

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heat exchanger, a throttle valve and an evaporator. From the t-h diagram, the characteristic of heat

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rejection at a supercritical pressure is a gliding temperature instead of a constant temperature at the gas

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cooler.

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Fig. 1 t-h diagram showing the transcritical CO2 cycle for water heating

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The system COP is calculated by Eq. (1).

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COP =

q h2 − h3 = w h2 − h1

(1)

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where q is the useful heating capacity per mass flow rate, [W]; w is the power input, [W]; h1 and h2 are

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the inlet and outlet refrigerant enthalpy values of the compressor respectively, h3 and h4 are the inlet

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and outlet refrigerant enthalpy values of the gas cooler, respectively.

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The throttle loss was defined as the enthalpy difference between the isoenthalpic throttling and isentropic throttle at the same evaporating temperature and is calculated by Eq. (2). 11

ACCEPTED MANUSCRIPT ∆h4' −5' = h4' − h5'

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(2)

where h4’, h5’ are the outlet enthalpy of the throttle vale at isoenthalpic and isentropic throttling processes, respectively [kJ/K•kg]. Fig. 2 illustrates the evaluation results of various inlet water temperatures in the gas cooler

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increased from 20 to 50ºC in a relative value to the COP and throttle loss of the theoretical cycle

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corresponding to the evaporating temperature of 0ºC and the discharge pressure of 12 MPa. Since the

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supply/return water temperatures are 75/50ºC in the heating system which terminal units is a radiator,

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considering the heat transfer temperature difference of 5ºC between the refrigerant and water as well as

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the expansion is a constant enthalpy process, i.e. the CO2 temperature at the exit of gas cooler was from

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25 to 55ºC, but other losses are neglect that means the compression is isentropic and there are no

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pressure drops in evaporator and gas cooler as well as pipeline. Therefore, it takes the inlet water

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temperature of 55ºC in gas cooler as the baseline in the theoretical cycle. In the figure, the theoretical

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cycle baseline COP and ∆h were 2.45 and 25.68 kJ/kg corresponding to the inlet water temperature of

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50ºC in gas cooler when the evaporating temperature was 0ºC. As the inlet water temperature in gas

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cooler decreased, the system COP increased linearly and the throttle loss decreased obviously.

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Compared with the baseline, the relative value of system COP and the throttle loss were 1.82 and 0.44,

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respectively, corresponding to the inlet water temperature of 25ºC in gas cooler. The reason for this

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phenomenon was that the refrigerant temperature at the exit of gas cooler increased with the increase of

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the inlet water temperature in gas cooler that caused, on the one hand, the entropy increased obviously

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during the throttle process leading to throttle loss increased. On the other hand, the enthalpy difference

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(∆h2-3) between the inlet and outlet of gas cooler declined due to the inlet water temperature in gas

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cooler rising, leading to a reduction in the heating capacity, while the compressor consumption (∆h2-1)

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was constant, that eventually resulted in the system COP decreased. As pointed out above, the higher inlet water temperature in the gas cooler caused a relatively high

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throttle loss and a lower system COP for CO2 heat pump during the heating mode. So decreasing the

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inlet water temperature in gas cooler is an effective method to reduce of throttling loss and improve the

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system performance. Therefore, in this paper, a sub-cooling technology with TES was adopted to

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decrease the inlet water temperature in the gas cooler to improve the system performance. On the other

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hand, using the energy stored in TES for space heating during night, which the ambient temperature

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was low that showed a lower system COP, to enhance energy utilization efficiency. The working

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principle of the CO2 heat pump heating system integrated with TES is described in detail in Section

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2.2.

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2.0

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Relative value

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Theoretical cycle baseline COP=2.45 Theoretical cycle baseline ∆h=25.86 kJ/kg

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0.44

0.49

COP ∆h

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Temperature (°C)

Fig. 2 Variation of relative values of COP and throttle loss (∆h) for different inlet water

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temperature in gas cooler

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2.2 Working principle of the system

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A detailed schematic diagram of a transcritical CO2 heat pump heating system integrated with TES

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is illustrated in Fig. 3(a). It consists of two close loops: the ASHP unit with CO2 as the working fluid, 13

ACCEPTED MANUSCRIPT and supply/return water close loop. For the CO2 heat pump unit, it consists of four main components,

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i.e. a reciprocating compressor (1), a gas cooler (2), electronic expansion valves (EEVs) (5), and an

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evaporator (6). Some additional devices such as intermediate heat exchanger (4), dry filter (3) and

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accumulate (7) are also required to ensure reliable operation of the system. For the supply/return water

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close loop, it consists of solenoid valves (8, 9,13), TES (L-low melt point PCM; H-high melt point

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PCM ) (10), a radiator (11) and a water pump (12). The system can operate in two different modes, i.e.

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heat charging mode (Fig. 3(b)) and heat discharge mode (Fig. 3(c)).

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(a) Schematic diagram

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(b) Heat charging mode

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(c) Heat discharging mode.

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Fig. 3 Schematic diagram of the transcritical CO2 heat pump heating system integrated with TES.

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In this paper, the main focus is on the supply/return water close loop. During the heat charging

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mode, the solenoid valves (8,9) are open while the solenoid valve (13) is close. Water is first heated by

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the CO2 ASHP unit, after the supply water flow is divided into two streams by a 3-way damper: one

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stream passes through the main pipe line and the rest is bypassed to the TES (10-H), where the two

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supply water streams are mixed before enters into the radiator (11) and releases the heat for space

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heating. The return water (the supply water after pass through radiator is regard as return water)

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passing through the water pump (12) and then flows through TES (10-L), the waste heat is absorbed by

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TES and resulting in a further reduction in the return water temperature. Finally, the return water enters

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into the ASHP unit. The refrigerant temperature before the throttle valve (5) is dropped due to the

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decrease in the temperature of the return water, leading to reduce the loss of throttling, resulting in an

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increase in the performance of the CO2 ASHP unit. During night, the performance of the CO2 ASHP

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unit decreases because of the low ambient temperature. In this situation, to improve the system

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efficiency, the ASHP unit is turned off and the heating system should be switched to the TES heat

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discharge mode.

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During the heat discharge mode, the ASHP unit is turned off, and the solenoid valve (13) is open 15

ACCEPTED MANUSCRIPT while the solenoid valves (8, 9) are closed. Firstly, the return water enters from the radiator (11) into the

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TES (10-L) to absorb heat, and it then flows through the TES (10-H) and the water temperature rises

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further that can meet the temperature requirement of supply water. Finally, the supply water flows

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through the radiator (11) and releases heat for space heating.

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2.3 Building model and boundary conditions

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The building concerned in this study was a typical single family rural house located in Beijing,

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China. The ground floor was 150 m2 and the total heated space area (i.e. three bedrooms and a living

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room) was 75 m2. The window to wall area ratio on the south and north facades of the house were 0.45

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and 0.2, respectively. The U-values of the building wall, floor and roof were 0.4, 0.039 W/m2•K and

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0.8W/m2•K, respectively. Windows had a U-value of 2.8 W/m2•K and a g-value of 0.5. The infiltration

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rate was set to 0.5 air changes per hour. The thermal transmittance values of the main building elements

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are present in Table 1. The indoor air temperature for the heating system design was 18°C to meet the

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requirement of indoor temperature and thermal comfort. The house model was created using TRNBuild

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(Type 56). To better understand the proposed system performance and to simplify the research of

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control influences, only a single zone building was defined in this paper, and thermal bridging was

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ignored for simplification.

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Table 1 Thermal transmittance values of main building elements

Building envelope

U-value

element

(W/m2•K)

1

Exterior wall

0.4

2

Roof

0.8

Cement +adhesive power polystyrene 100 mm;

3

Ground floor

0.039

Concrete slab100 mm

4

Windows (g=0.5)

2.8

Double lay plastic steel window;

NO.

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Structure Concrete brick 370mm; polystyrcne thermal insulating board50 mm

2.4 CO2 heat pump model The investigation started with an evaluation of an air-to-water CO2 heat pump unit. The nominal 16

ACCEPTED MANUSCRIPT heating capacity and COP of the baseline unit were 8kW and 4, respectively. This baseline system was

323

then compared to the proposed system integrated with PCMs. In TRNSYS, the air-to-water heat pump

324

(i.e. Type 941) was modelled using a performance map with the data of heating capacity and power

325

consumption, which were the functions of the ambient temperature and inlet water temperature in the

326

gas cooler, for a range of testing points, i.e. the heat output and COP of the heat pump were calculated

327

in the model through interpolation between these points.

SC

329

The technical characteristics of the CO2 heat pump unit were provided by the manufacturer. The heating capacity (Eq. 3) and power consumption (Eq. 4) can be fitted as follows:

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322

330

Qmax = 19.02 + 0.402a − 0.5Tin + 0.00112Ta2 − 0.0088Ta • Tin + 0.0012Tin2 −

331

Pmax = 3.35 + 0.076Ta + 0.12Tin + 6.14 × 10−4 Ta2 − 0.00195Ta • Tin − 0.003Tin2 +

2.3 × 10−5 Ta3 − 3.06 × 10−5 Ta2 • Tin + 8.59 × 10−5 Ta • Tin2 − 9.56 × 10−5 Tin3

2.29 × 10−6 Ta3 − 1.57 × 10−5 Ta2 • Tin + 1.9 × 10−5 Ta • Tin2 + 2.62 × 10−5 Tin3

(3)

(4)

where Qmax and Pmax are the maximum CO2 heat pump unit heating capacity and power input

333

respectively, [kW]; Ta is the ambient temperature, [°C]; and Tin is the inlet water temperature in the gas

334

cooler, [°C].

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332

Air-to-water CO2 heat pump unit taken act air energy as low resource, the evaporator is inevitably

336

subject to frosting which results in a reduction in the heating capacity and system performance when it

337

operates in winter. The previous studies showed that the system heating capacity and COP decreased by

338

30 ~ 57% and 35 ~ 60%, respectively, when frosting occurred on the evaporator [33][34][35]. To

339

ensure the system run safely and reliably, periodical defrosting is therefore necessary. Literatures

340

[36][37] presented a COP reduction (COPreddefrost) which defined by a modified gaussian curve derived

341

via the application of a biquadratic polynomial approximation. It indicated that the defrost parameter

342

can be express as follows when the ambient temperature (Ta) is below 7°C.

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17

ACCEPTED MANUSCRIPT 343

COPreddefrost = −0.0027 × (Ta − 7 ) + 0.1801 • e −Ta

344

When the ambient temperature is above 7°C, the defrost parameter can be given as:

COPreddefrost = 0.1801 • e − Ta

345

2

2

/5

(5)

/5

(6)

Therefore, the correct system (COPcorr), heating capacity (Qcorr) and water temperature at the exit of the

347

gas cooler (Tout) can be calculated as follows[38]:

COPcorr = COPicycle (1 − COPredefrost )

348

Qcorr =COPcorr • Pmax

350

Tout = Tin − Q / ( c • m )

(7) (8)

(9)

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349

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346

where COPicycle is the system COP which was not taken into account the effect of frosting; Pmax is the

352

ASHP unit maximum thermal output power, [kW]; and Tin is the inlet water temperature of the gas

353

cooler, [°C].

354

2.5 Heat transfer model of TES

355 356

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351

TES with PCMs is an effective energy storage method. Numerous research indicated that PCMs have high energy storage densities and are ideal candidates for cold/heat storage [39]. In the present study, a cylindrical tank (tube-in-tank arrangement) filled with PCMs as the TES

358

was considered and the mathematical model was formulated using ε −  method based on Tay’s

359

[40] research.

361 362 363 364

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357

The instantaneous effectiveness of the heat exchanger is defined by heat flow to a heat source/sink

of infinite specific heat as shown in Eq. (10).

ε =1-exp ( − NTU )

(10)

where NTU is the number of transfer units determined by Eq. (11).

NTU =

UA 1 = mC p RT mC p 18

(11)

ACCEPTED MANUSCRIPT 365

where U is the overall heat-transfer coefficient, [W/(m2•K)]; A is the heat transfer area, [m2]; m is the

366

mass flow rate of the heat transfer fluid, [kg/s]; Cp is the specific heat of the heat transfer fluid,

367

[kJ/(kg•K)]; and RT is the total thermal resistance, [K/W]. ln ( Ro / Ri ) 1 1 + + 2π Ri Lh f 2π k w L Sk PCM

(12)

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RT =

368

where Ri is the inner radius of the tube, [m]; Ro is the outer radius of the tube, [m]; L is the length of the

370

tube (m); hf is the heat transfer coefficient of the heat transfer fluid, [W/(m2•K)]; kw is the thermal

371

conductivity of the tube wall [W/(m•K)]; kPCM is the thermal conductivity of the PCM (W/(m•K)); and

372

S is the shape factor of the PCM [m], which is determined using Eq. (13).

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369

373

375

follows:

δ=

(14)

shown in Eq. (15).

•

ε =1- exp  −1 / m Cp RT  



(15)

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379 380

R2 − Ro2 2 Rmax − Ro2

According to Eqs. (10) – (14), the heat exchanger effectiveness can be written in terms of RT as

EP

378

(13)

The phase change fraction (δ) for a tube surrounded by a cylindrical volume of PCM is defined as

376

377

2π L ln( R / Ro )

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374

S=

The heat transfer coefficient (hf) of the HTF (W/(m2•K)) is presented by Eq. (16) [41], in which

381

the Nusselt number for the laminar flow and turbulent flow was determined using Eq. (17) and Eq. (18),

382

respectively.

383

384

h f = ( Nuk f ) / d i Nu = 3.66 +

[0.0668(d i / L) RePr ] 1 + 0.04[( d i / L ) RePr]2/3

385 19

(16)

(17)

ACCEPTED MANUSCRIPT Nu = 0.023Re0.8 Pr n .

386 where n = 0.3 and 0.4 for cooling and heating.

388 (20).

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389

The Reynolds number is determined using Eq. (19) and the Prandtl number is determined by Eq.

•

390

m di Re = Ac µ f

391

Pr =

20 15 10 5 0

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Temperature (°C)

55

50

40

EP

45

100

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0

395

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Experimental Simulated Relative error

60

394

(20)

kf

where µ f is the dynamic viscosity of the heat transfer fluid, [kg/(m•s)].

65

393

µ f Cp

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(19)

-5

Relative error (%)

387

(18)

-10 -15 -20

200

300

400

Time (minutes)

Fig. 4 Comparison between simulated and experimental data.

The performance of this model was validated using the experimental data presented by Li et al.

396

[42]. The PCM used in the experiments was Paraffin (the thermal properties are shown in Table 2) and

397

the inlet water temperature was 60ºC. It can be observed from Fig. 4, the simulation data generally

398

matched with the experimental data and the maximum relative error between the simulated data and

399

experimental data was 7.15%. The precision of the entire simulated results was within the acceptable 20

ACCEPTED MANUSCRIPT 400

accuracy level of simulation for PCM analysis, which showed that it can be used to analyze the heating

401

characteristic of the proposed system.

402

Table 2. Thermal physical parameters of PCM Value

Density(kg/m3)

916(293.15K), 776(343.15K)

Specific heat load(kJ/(kg·K))

1.7(293.15K), 2.5(343.15K)

Phase change temperature(K)

319.83

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Name

0.27

SC

Thermal conductivity(W/(m· K)) Latent heat(kJ/kg)

141.91

404

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403 2.6 Operation modes and control strategies

The model of the proposed system was established in TRNSYS 17.0, which is a modular

406

simulation software that can be effective to evaluate and optimize the behavior and performance of

407

transient thermal systems. Fig. 5 illustrates the simulation system developed, in which the majority

408

component models used were the standard models provided in the TRNSYS library. The TES model

409

was developed according to the key governing equations proposed by Tay et al. [40].

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405

For the annual heating performance of the proposed system, the simulation period started from

411

November 15th to March 15th of the next year, which is the winter space heating period in Beijing, with

412

an indoor set-point temperature of 18°C.

413

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410

The control strategy for heat pump systems is very important to improve its performance. In this

414

paper, an on/off (type 2b) controller was employed to model the practical operation of the heat pump

415

and water pump subjected to the temperature control. The upper input temperature (Th) was set to 21°C,

416

and the lower input temperature (Tl) was set to the indoor real-time temperature. The upper and lower

417

dead bands of the differential temperature controller were set to 5 and 1°C, respectively. According to 21

ACCEPTED MANUSCRIPT 418

the start and stop conditions of the ASHP and the heat supply, the target was to maintain to the indoor

419

temperature to be 18 ± 2°C.

Turn Irradiation

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Radiation Shading+Light

Weather data

Lights

Light Thresholds

SC

Building

Temperature

Raditor

ASHP

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Water pump

Controler

420 421

Average

performance

Turn

Radiation

Weather data

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(a) The baseline system

Irradiation

Shading+Light

Lights

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Light Thresholds

Flow diverter-2

Building

Temperature output

Water pump

PCM_L

Raditor

ASHP

Controler

Time

Performance output

Flow mixier-2

PCM_H

Equa

422 423 424

Flow diverter

Flow mixer

(b)The proposed system Fig. 5. TRNSYS model of the baseline and proposed heating system. 22

ACCEPTED MANUSCRIPT Due to the fact that the ambient temperature was relative low at night which led to a reduction in

426

the system heating capacity and COP, in order to improve the heating system performance in over

427

heating period, in this paper, it was assuming that the electricity price was constant during the day and

428

night, the CO2 heat pump unit was operated for heating from 6:00 to 22:00, and the unit was turned off

429

from 22:00 to the next day of 6:00 and heated by the TES. Table 3 shows the main TRNSYS

430

components and parameters settings, and the logical running control of the proposed systems is

431

sketched in Fig. 6.

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432

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425

Table 3. Main TRNSYS components and parameters setting. Component type

Main parameters

Building

Type 56

Radiator

Type 1231

ASHP

Type 941

Pump

Type 3d

Flow diverter

Type 11f

Control signal:0.4

Weather data

Type 15

Input Beijing typical meteorological years data

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Name

Room air exchange rate: 0.5h-1;

Controlled temperature: 18±2°C Design capacity 5 kW;

Design surface temperature: 55°C. Rated heating capacity: 8 kW; Rated heating power: 2.3 kW Maximum power: 0.067kW

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Maximum flow rate 300kg/h;

433 434

Fig. 6 Logical diagram of the system operation control. 23

ACCEPTED MANUSCRIPT 435

3. Results and discussion In this section, the operating characters of the CO2 heat pump heating system integrated with TES

437

during an entire heating period was analyzed through the TRNSYS 17.0 software with 10minutes time

438

step. In order to verify the feasibility of the proposed system, the primary performance including room

439

temperature, system COP, heating capacity and energy consumption as well as HSPF (Heating

440

Seasonal Performance Factor) was compared with a conventional CO2 heat pump system without TES.

441

3.1 Comparison on indoor temperature

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436

The relevant simulation outputs including the indoor temperature (Tindoor represents the proposed

443

system; Tindoor-base represents the baseline system) during the entire heating period are shown in Fig. 7. It

444

can be seen that most of the indoor temperature varied between 16 and 20ºC which met the requirement

445

of the occupant comfort levels for the both of the system in the entire heating period. However, for the

446

conventional CO2 heat pump system, the indoor temperature was dropped a few degrees below than the

447

set-point from Jan. 10 to 18 and Jan. 22 and 23, which the indoor temperature was around from 9.8 to

448

16ºC that was seen that the both system could not meet the set-point temperature and required backup

449

heating in January. This is caused by the fact that the lower ambient temperature (Tamb) and higher

450

heating loads of the house in January during the entire heating period in Beijing, i.e., the coldest month,

451

leading to the decrease in the supply water temperature (see Fig. 8), and further led to the decrease in

452

the indoor temperature. It should be noted that, for the proposed system, the Tindoor was always maintain

453

at 16 ~ 20ºC in January (see the right of Fig. 7). This illustrated that the proposed system in the indoor

454

thermal comfort was better than the baseline system.

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During the entire heating period, as the ambient temperature in the January was the lowest in

456

comparison with the other months, the following analysis was mainly focusing on the characteristic of 24

ACCEPTED MANUSCRIPT 457

the proposed system in the coldest month including supply/return water temperature, heating capacity

458

and unit COP.

Tindoor-base，

-10

Nov 30

Dec 15

Air temperature (°C)

Dec 30

Jan 14

Date Tindoor-base，

20 18 16 14 12 10 Jan 7

Jan 29

Feb 13

Jan 13

Feb 28

Mar 15

SC

-20 Nov 15

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0

Jan 1

Tindoor

Jan 19

Jan 25

Jan 31

Date

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459

461

Tamb See the below figure

10

22

460

Tindoor，

20

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Air temperature (°C)

30

Fig. 7 Comparison of room temperature for the proposed system and baseline system.

3.2 Comparison on supply/return water temperature of the CO2 heat pump unit Fig. 8 shows the variations of supply/return water temperatures of the CO2 heat pump unit for the

463

proposed system and baseline system in January during the heating period. For the conventional CO2

464

heat pump system, the supply/return water temperature decreased at the beginning, reaching the

465

minimum, and then increased. The maximum and minimum supply water temperatures were 79.5 and

466

39.6ºC respectively, corresponding to the return water temperatures of 66.2 and 33.2ºC. It can also be

467

seen that the temperature difference between the supply and return water was around 7 ~ 10ºC. During

468

the heating period from 12th to 17th Jan 12th to 17th and around Jan. 24th, the supply/return water

469

temperature reached the lowest value. This was because the decrease in ambient temperature caused the

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25

ACCEPTED MANUSCRIPT reduction in the evaporation temperature of the ASHP, leading to a decrease in the heating capacity,

471

resulting in the decrease in the supply water temperature. In addition, frost may be accumulated on the

472

surface of the outdoor heat exchanger which increased the heat transfer resistance between the outdoor

473

air and heat exchanger, so the CO2 heat pump unit needs to use periodic frost-defrost cycle, which

474

further increased this phenomenon.

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470

For the proposed system, the supply water temperature reached the peak value of 72.1ºC and the

476

vale value of 48.3ºC, and the maximum and minimum return water temperatures were 58 and 33.3ºC.

477

Compared to the baseline system, the variation of the average supply/return water temperature per day

478

showed a curve gradually. In addition, the temperature of the water return to heat pump was relative

479

lower (average value: 36.5ºC) than 16.8ºC of the baseline system (average value: 53.3ºC), and the

480

temperature difference of the supply/return water was relative large which was around of 16 ~ 20ºC.

481

This was because the return water from radiator first entered to TES, and much heat was absorbed by

482

the PCM. According to the previous theory analysis (Fig. 2), the throttle loss of the proposed system

483

was reduced due to the decrease in the low return water temperature. Moreover, the heat stored in the

484

TES for heating at time of 22:00 to 6:00, that the ambient temperature was very low caused a poor

485

performance of the ASHP system, both of these reasons resulted in the increase in the performance of

486

CO2 heat pump unit (also can be seen in Fig.10).

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487

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475

26

80 70

Tsupply,

Taverage supply

Treturn,

Taverage return

Baseline system

60 50 40 30 Jan 1

Jan 7

Jan 13

Jan 19

Tsupply,

Taverage supply

70

Treturn,

Taverage supply

30 Jan 1

Jan 7

Jan 13

Jan 19

Jan 25

Jan 31

Date

Fig. 8 Comparison of supply/return water temperatures for the proposed system and baseline

490

system.

3.3 Comparison on system heating capacity

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491

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40

488 489

Proposed system

60 50

Jan 31

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Water temperature (°C)

Date 80

Jan 25

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Water temperature (°C)

ACCEPTED MANUSCRIPT

Fig. 9 shows the simulation results for the system heating capacity for the proposed system and

493

baseline systems in January. From this figure, it can be seen that the heating capacity of the proposed

494

system was 2013 kWh, which was 22% lower in comparison with the baseline system of 2581 kWh.

495

The reason was that the proposed system was operated in days that the ambient temperature was much

496

higher than night, leading to much higher heat pump energy efficiency coefficient. In addition, more

497

energy was stored in TES for heating space at night, which reduced the time of frosting accumulated on

498

the surface of the heat exchanger, and further improved the efficiency of the proposed system and

499

decreased the heating energy consumption (Table 4).

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500

27

ACCEPTED MANUSCRIPT

Jan 7

Jan 13

503

Jan 19

Date

8 7 6 5 4 3 2 1 0 Jan 1

Studied system

Qtotal=2013 kWh

Jan 25

SC

Jan 7

Jan 13

Jan 19

Jan 31

Jan 25

Jan 31

Date

501 502

Qtotal-base=2581 kWh

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Baseline system

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Heating capacity (kW)

Heating capacity (kW)

8 7 6 5 4 3 2 1 0 Jan 1

Fig. 9 Comparison of heating capacity of the baseline system and proposed system.

3.4 Comparison on COP of the CO2 heat pump unit

Fig. 10 shows the variations of the COP of the CO2 heat pump unit for the proposed system and

505

baseline system in January. It can be seen that the variation trends of the COP of the both systems were

506

similar, which first decreased with the increase in the heating time and then increased after reaching the

507

minimum value at the time from the January 12th to 17th. The phenomenon for this change was due to

508

the increase and decrease in ambient temperature, and the details have been presented in 2.3 section.

509

The maximum, minimum and average COP values of the CO2 heat pump for the proposed system and

510

the baseline system were 3.78, 1.44, 2.39 and 3.59, 1.25, 2.23, respectively, in January. The average

511

value of COP was 7.2% higher than the baseline system. This illustrated that the performance of the

512

proposed system, which used TES to absorb the waste heat from the return water and for heating at

513

night, was outperformed that of the baseline system. It can be observed that the system COP was

514

significantly influenced by the gas cooler inlet temperature. For instance, the COP increased with the

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504

28

ACCEPTED MANUSCRIPT decrease in the return water temperature. 4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0 Jan 1

COPaverage-base=2.23

Baseline system

Jan 7

Jan 13

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COP

515

Jan 19

Jan 25

4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0 Jan 1

Jan 31

COPaverage= 2.39

SC

Proposed system

M AN U

COP

Date

Jan 7

Jan 13

Jan 19

Jan 25

Jan 31

Date

516

Fig. 10 Comparison of COP values of the CO2 heat pump for the baseline system and proposed system.

518

3.5 Comparison on the characteristic of the both system in the entire heating period.

519 520

Table 4. Summary of simulation results of the proposed system and baseline system during the entire

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517

Month

Heating capacity

Energy consumption

(kWh)

(kWh)

EP

heating period

System COP (HSPF)

Base

Proposed

Base

Proposed

Base

system

system

system

system

system

system

Nov.

953

1166

389

505

2.45

2.31

Dec.

1673

2212

725

971

2.31

2.28

Jan.

2013

2581

887

1195

2.27

2.16

Feb.

1659

2033

679

860

2.44

2.36

Mar.

723

876

288

361

2.51

2.43

7021

8868

2968

3892

(HSPF)

(HSPF)

2.37

2.28

AC C

Proposed

Entire heating period (kWh) Comparison with base system (%)

-0.21

-0.24

521 29

0.04

ACCEPTED MANUSCRIPT The overall results including the heating capacity, power energy and HSPF (heating seasonal

523

performance factor) of the baseline system and the proposed system are summarized in Table 4. The

524

HSPF is an important factor to evaluate the heating performance of an ASHP unit. The higher the

525

HSPF rating of a unit, the more energy efficiency it is. The HSPF, defined as a ratio of heating over the

526

heating season to power consumption, as expressed in Eq. (21). For the proposed system during the

527

entire heating period, it provided 7021 kWh of heating while consumed 2968 kWh of electricity which

528

including heat pump (compressor + control + blower) and water pump power consumption.

529

Therefore, the HSPF turned out to be 2.37. Compare with the baseline system, the heating capacity

530

(8868kWh) and power energy (3892 kWh) decreased by21 and 24%, respectively. However, the HSPF

531

of the proposed system was increased by 4% in comparison with the baseline system (2.28).

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522

τ

∫ Qdτ HSPF = τ ∫ Pdτ 0

532

(21)

0

where Q is the total heating capacity, [kWh]; and P is the total energy consumption including heat

534

pump (compressor + control + blower) and water pump, [kWh].

TE D

533

From this table, it can also be seen that the largest and lowest monthly system COPs were

536

occurred in March and January, respectively, when the average ambient temperature of 5.1 and -4.3°C.

537

The system COP of the baseline system and the proposed system were 2.16 and 2.27 respectively in

538

January, and were 2.43 and 2.51, respectively, in March. This demonstrated that the heating unit

539

performed poorly in January (the coldest month) and the indoor temperature was below the set-point

540

temperature in the zone.

541

4. Conclusions

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542

A study of a transcritical CO2 heat pump heating system integrated with TES for a rural

543

single-family in heating period in Beijing, China, using TRNSYS simulation has been carried out. The 30

ACCEPTED MANUSCRIPT TES model was formulated using the  −  method and the simulated results showed a good

545

agreement with the experimental data. The simulation results were compared in terms of operation

546

characteristics with a convention transcritical CO2 heat pump heating system without TES during the

547

entire heating season. The main conclusions drawn from the results are as follows.

548

(1)

RI PT

544

Taking the inlet water temperature of 50ºC in the gas cooler at the evaporating temperature of 0ºC as the baseline cycle corresponding to the system COP and ∆h were

550

2.45 and 25.68 kJ/kg, the relative value of system COP and the throttle loss increased to

551

1.82 and decreased to 0.44 when the inlet water temperature decreased to 25ºC in gas

552

cooler. (2)

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553

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549

The indoor temperature during the heating period by using the proposed system can be maintained at 16 ~ 20ºC in January, but for the conventional system, the indoor

555

temperature was around from 9.8 to 16ºC during Jan. 10 to 18 and Jan. 22 and 23. This

556

illustrated that the proposed system provided a better indoor thermal comfort than the

557

baseline system. The average temperature of the water return to the CO2 heat pump unit

558

was 36.5ºC, which was 16.8ºC lower than the baseline system with the average value of

559

53.3ºC.

561 562

EP

AC C

560

TE D

554

(3)

The heating capacity and energy consumption of the proposed system were 8868 and 2968 kWh, respectively, during the entire heating period, which were decreased by 21 and 24%, respectively, in comparison with the baseline system. The HSPF of the

563

proposed system was 2.37 during the entire heating period, which was increased by 4%

564

when compared with the baseline system.

565

Acknowledgements 31

ACCEPTED MANUSCRIPT The work was supported by National Natural Science Foundation of China (No. 51606139), China

567

Postdoctoral Science Foundation Funded Projects (No. 2016M590950 and No. 2017T100753), Shaanxi

568

Province Postdoctoral Science Foundation Funded Projects (No. 2017BSHTDZZ17) and Xi’an

569

Municipal Science and Technology Projects (No. 2017078CG/RC041(XAJD001).

570

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[2] R. Yao, B. Li, K. Steemers. Energy policy and standard for built environment in China, Renew. Energy 30 (2005) 1973–88.

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[3] Y. Wang, F. Wang, H. Wang, Heating energy consumption questionnaire and statistical analysis of rural buildings in China, Procedia Engineering, 146(2016)380-385.

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[5] Z.H. Wang, M.J. Song, F.H Wang, et al., Experimental investigation and seasonal performance assessment of a frost-free ASHP system with radiant floor heating, Energy Build. 179(2018) 200-212

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[6] K.J. Chua, S.K. Chou, W.M. Yang, Advances in heat pump systems: a review, Appl. Energy 87 (2010) 3611–3624.

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[7] Z.H. Wang, F.H. Wang, Z.J. Ma, et al. Field test and numerical investigation on the heat transfer characteristics and optimal design of the heat exchangers of a deep borehole ground source heat pump system, Energy Convers. Manag. 153 (2017) 603–615.

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Figure Captions Fig. 1 t-h diagram showing the transcritical CO2 cycle for water heating. Fig. 2 Variation of relative values of COP and throttle loss (∆h) for different inlet water temperature in

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gas cooler Fig. 3a Schematic diagram Fig. 3b Heat charging mode

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Fig. 3c Heat discharging mode.

Fig. 3 Schematic diagram of the transcritical CO2 heat pump heating system integrated with TES.

Fig. 5a The baseline system Fig. 5b The proposed system

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Fig. 4 Comparison between simulated and experimental data.

Fig. 5.TRNSYS model of the baseline and proposed heating system.

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Fig. 6 Logical diagram of the system operation control. Fig. 7 Comparison of room temperature for the proposed system and baseline system.

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Fig. 8 Comparison of supply/return water temperatures for the proposed system and baseline system. Fig. 9 Comparison of heating capacity of the baseline system and proposed system.

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Fig. 10 Comparison of COP values of the CO2 heat pump for the baseline system and proposed system.

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ACCEPTED MANUSCRIPT Table 2. Thermal physical parameters of PCM Value

Density(kg/m3)

916(293.15K), 776(343.15K)

Specific heat load(kJ/(kg·K))

1.7(293.15K), 2.5(343.15K)

Phase change temperature(K)

319.83

Thermal conductivity(W/(m·K))

0.27

141.91

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Latent heat(kJ/kg)

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Name

ACCEPTED MANUSCRIPT Highlights: A CO2 heat pump heating system integrated with thermal energy storage was proposed.




The heating characters of the system were studied using the TRNSYS simulation.




The HSPF of the system increased by 4% in the entire heating period.




A better system with a significant energy saving was achieved.

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