Experimental investigations of an absorption heat pump prototype with intermediate process for residential district heating

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Experimental investigations of an absorption heat pump prototype with intermediate process for residential district heating ⁎

Ding Lua,b, Yin Baia,b, Yanxing Zhaoa,b, Xueqiang Donga,b, Maoqiong Gonga,b, , Ercang Luoa,b, Gaofei Chena, Qingyu Xua,b, Jun Shena,b a b

Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, Beijing 100190, China University of Chinese Academy of Sciences, Beijing 100049, China

A R T I C LE I N FO

A B S T R A C T

Keywords: Absorption heat pump District heating Natural gas Experimental study Waste heat recovery Intermediate process

Despite being efficient means of district heating, conventional single-effect absorption heat pumps suffer significant performance deterioration at low ambient temperatures. To solve this problem, an ammonia-water absorption heat pump prototype with intermediate process is designed and built. It utilizes both low-grade heat from the ambient and exhaust heat from natural gas combustion, through the evaporator and intermediate evaporator, respectively. To perform comparison study, the prototype also has the single-effect mode. Experimental results indicate that when the evaporation temperature is 0 °C, the prototype with intermediate process can provide 30 kW heating capacity to heat the water in sequence through the rectifier, condenser and absorber, from 34.24 to 55.09 °C, and the coefficient of performance and primary energy efficiency are 1.66 and 1.28, respectively. When evaporation temperatures reduce to −5, −10 and −15 °C, the coefficient of performances are 1.51, 1.40 and 1.28, respectively, reaching 77%–90% of the simulation values under corresponding working conditions. Compared to the conventional single-effect system, the prototype with intermediate process performs better at lower evaporation temperature, with the coefficient of performance improved by 11%. By replacing the original heat exchangers, the proposed system can be easily applied to the existing heat networks based on gas-fired boilers to improve energy efficiency, and the favorable experimental performance proves that it is a more efficient way of residential district heating, especially in cold regions.

1. Introduction In China, buildings consume a great portion of energy, accounting for nearly 24% of the total energy consumption [1], and this number will gain further increase considering the rapid urbanization and gradually improved living environment [2]. A huge amount of the building energy is utilized by residential district heating systems, and only in winter in northern China, they occupy more than 20% [3]. At present, the most common district heating methods are coal- and gas-fired boilers [4], which have two main drawbacks: causing severe air pollution and having low energy efficiency. Therefore, many novel technologies [5] are developed to partly replace these boilers, and the absorption heat pump (AHP) [6] stands out among them, on account of its higher primary energy efficiency (PEE) [7]. Many AHP district heating systems use natural gas combustors or heat networks as the driving source, and extract low-grade heat from either the air [8] or the ground [9]. By recovering additional ambient heat, the PEE exceeds 1, and the energy saving potential is 20% [10] compared to conventional boilers.

⁎

Single-effect AHP systems applied for district heating were widely studied theoretically and experimentally. Some literatures focused the systems driven by heat networks, and a ground-source AHP cycle [11] for district heating was proposed, and was shown to be more efficient in energy utilization compared to electrically-driven heat pumps. Further studies [12] indicated that compare to conventional heating systems based on boilers, the proposed system improved the energy efficiency by 54%. Wu et al. [13] analyzed a water-source AHP cycle producing 45 °C hot water for low-temperature district heating. Experimental study showed that when the evaporator inlet temperature is −10 °C, the coefficient of performance (COP) could reach 1.2. Nitkiewicz et al. [14] compared the performance of an electrically-driven heat pump, an AHP, and a gas-fired boiler, and demonstrated that the heating plant using ground source had a much lower eco-indicator than using a gasfired boiler, and the AHP had a lower environmental impact than the electrically-driven heat pump. Li et al. [15] applied an AHP system to recover heat from a municipal heat network, and it was indicated that by utilizing low-grade renewable energy, both the heating capacity and

Corresponding author at: Key Laboratory of Cryogenics, Technical Institute of Physics and Chemistry, Chinese Academy of Sciences, Beijing 100190, China. E-mail address: [email protected] (M. Gong).

https://doi.org/10.1016/j.enconman.2019.112323 Received 27 September 2019; Received in revised form 12 November 2019; Accepted 19 November 2019 0196-8904/ © 2019 Elsevier Ltd. All rights reserved.

Please cite this article as: Ding Lu, et al., Energy Conversion and Management, https://doi.org/10.1016/j.enconman.2019.112323

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Nomenclature

cP m P PM Q RWH T ΔT V W x ηC ηE ρ ωR ωS ωT

COP GAX HACHP PEE

specific heat,kJ/(kg·°C) mass flow rate, kg/h pressure, MPa intermediate pressure, MPa heat duty, kW ratio of waste heat temperature, °C temperature difference, °C volume flow rate, m3/h work, kW concentration efficiency of the combustor efficiency of electricity generation density, kg/m3 random uncertainty system uncertainty total uncertainty

Subscripts A C E G in IA IE O out P R RW SS SW W WH

Abbreviations AHP

coefficient of performance generator-absorber heat exchange hybrid absorption-compression heat pump primary energy efficiency

absorption heat pump

absorber condenser evaporator generator inlet intermediate absorber intermediate evaporator oil outlet solution pump rectifier return water strong solution supply water water waste heat

temperature of −18.3 °C. Besides this advantage, Farshi et al. [24] also found that the NH3–LiNO3 system had better performance under lower generation temperature, meanwhile requiring no rectifier. For improving system structures, the main method is introducing compressor to the single-effect AHP system, that is the hybrid absorption-compression heat pump (HACHP). The compressor is commonly placed between the evaporator and absorber to increase the absorption pressure under low evaporation temperatures [25], or between the generator and condenser to increase the heating temperature [26]. Wu et al. performed experimental analysis on a HACHP system, and found that that the lower limit of the evaporation temperature could be extended from −10 to −25 °C by introducing the compressor [27]. They also proposed another form of HACHP system with a compressor in parallel with the generator and absorber [28], and simulation results indicate that increasing the ratio of refrigerant flow distributed to the absorption sub-cycle could improve system performance under lower ambient temperature. Liu et al. [29] carried out studies on both working fluids selection and system structure modification. Three different working pairs were applied to two types of HACHP systems. Results indicated that compared to the single-effect system, the HACHP systems could improve the performance including raising COP and enlarging operating range of the evaporation temperature. Moreover, placing the compressor between the evaporator and absorber performed better than between the generator and condenser. Besides the HACHP system, some literatures [30] proposed the double-stage absorption system to lower the evaporation temperature and improve system performance under low ambient temperatures. These studies usually involved new working fluids selecting, mainly the NH3–salts [31]. Moreover, there was another simple but useful method. Li et al. [32] proposed a district heating system where the air-source AHP was in parallel with a conventional plate heat exchanger, and when the evaporation temperature was too low, the plate heat exchanger started operating to ensure reliability of heat supply. Besides the performance deterioration in low evaporation temperatures, the other main drawback of the conventional single-effect AHP system is that, it lacks efficient means to recover flue gas waste heat, when it is either directly gas-fired or driven by a heating network based on the gas boiler (combustor). The flue gas generated from the natural gas combustion usually has the temperature higher than 150 °C [33], and contains a great deal of water vapor. Its latent heat occupies

PEE were improved. Moreover, the operating cost was reduced significantly resulting from the decrease of coal and power consumption. Some other literatures focused on the systems driven by directly combustion of natural gas. Keinath et al. [6] investigated a gas-fired AHP system that could heat the water to 57 °C. Theoretical analysis indicated that a COP of 1.74 was acquired when the ambient temperature was 20 °C. Under the same ambient temperature, the gas-fired AHP prototype proposed by Garrabrant et al. [16] had the measured COP of 1.63 when producing 45 °C hot water. Baig et. al. [17] developed and implemented a monitoring system to test the performance of a gas-fired AHP cycle. Results showed that the PEE kept above 1 when the ambient temperature was −5 °C, indicating that it is more efficient than conventional gas-fired boilers. However, when the ambient temperature was lower than −5 °C, the PEE reduced below 1. The above analyzed systems were all in single-effect forms. Despite being an efficient means of district heating, conventional single-effect AHP systems suffered significant performance deterioration at low ambient temperatures, limiting their application in cold regions. The performance deterioration is more serious for a complex system such as the generator-absorber heat exchange (GAX) system. Garimella et al. [18] found that when the ambient temperature reduced from 5.6 to −30 °C, the COP of a GAX cycle dropped sharply from 1.4 to 1.05. It was because when the ambient temperature fell down, the generatorabsorber temperature overlap reduced or even disappeared. To solve this problem, two major technical solutions were developed: using alternative working fluids [19], and improving system structures. For using alternative working fluids, some literatures suggested that the binary working fluids ammonia-lithium bromide (NH3–LiNO3) and ammonia-sodium thiocyanate (NH3–NaSCN) were excellent choices [20]. Cai et al. [21] carried out experimental analysis on the thermodynamic performance of a single-effect absorption system using NH3–LiNO3 and NH3–NaSCN as working pairs. Experimental results showed that the COP of the NH3–NaSCN system were higher than that of the NH3–LiNO3 system, while the NH3–LiNO3 system could operate under a lower evaporation temperature. Similar conclusion was drawn by Abdulateef et al. [22] They compared the performance of an absorption system using different working pairs, and found that the NH3–NaSCN system could hardly work when the evaporation temperature reduced below −10 °C. Furthermore, Táboas et al. [23] reported that the NH3–LiNO3 system could reach low evaporation 2

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In this paper, an experimental prototype based on the previous theoretical analysis [42] is designed and built. The prototype is driven by an electrical oil heater, which simulates the natural gas combustor, and heat conduction oil is used as the heating medium. It has two heating modes: the single-effect mode and intermediate-process mode. In the intermediate-process mode, the system utilizes both low-grade heat from the ambient and exhaust heat from natural gas combustion, through the evaporator and intermediate evaporator, respectively. In the single-effect mode, the solution preheater and intermediate evaporator do not transfer any heat, and the intermediate absorber changes into a conventional solution heat exchanger. Experimental investigations are carried out to study its performance in these two heating modes, and influences of various internal and external parameters are further analyzed, in order to understand how the prototype operates under different working conditions.

more than 10% of the total calorific value of natural gas [34]. Therefore, it is essential to fully recover the waste heat to improve the energy efficiency [35]. In most gas-fired heating systems, the flue gas waste heat is recovered by the return water through a condensing heat exchanger [36]. However, it was indicated that the return water temperature was quite similar to the water vapor dew point, and the latent heat could not be totally recovered [37]. To solve this problem, some other methods were proposed [38], including applying both the gasliquid heat exchanger and AHP system [39], which recovered the sensible and latent heat of the flue gas, respectively. Simulation results showed that the energy efficiency was improved by 10%. Instead of using AHP system to recover waste heat through the absorber and produce hot water, Zhu et al. [40] used the AHP to produce 20 °C cooling water to recover the flue gas waste heat, and the heating capacity was improved by 12% after cooling the flue gas down to 30 °C. Besides, an open-cycle AHP system [41] was applied and the PEE was increased by 1.5%. However, PEEs of the systems mentioned in this paragraph were all lower than 1, because the return water was mainly heated by natural gas combustion, while the AHP systems were only used to recover waste heat and produce small amounts of heating capacity. In order to solve the above two major problems, a gas-fired absorption heat pump for district heating in cold regions was proposed and analyzed theoretically in the previous study [42]. Intermediate evaporation and absorption processes [43] were introduced to improve the system performance at low ambient temperatures, and the sensible and latent heat of the flue gas was cascade recovered in the solution preheater and intermediate evaporator to enhance the waste heat recovery. Previous simulation analysis demonstrated that the PEE could exceed 1 even when the evaporation temperature was below −15 °C. Moreover, compared to the single-effect system, the proposed system had 13% improvement in COP, and the energy saving potential was 11.7%.

2. Design of the system The district heating network, based on the AHP system with intermediate process, is shown in Fig. 1. The driving heat is generated by natural gas combustion, and heat conduction oil is used as the heating medium, transforming heat from the combustor to the generator. Compared to direct-fired AHP, this form has a major advantage: it can be easily applied to an existing heat network based on the gas-fired boiler, only by replacing the original heat exchanger. Driven by the generation heat and pump work, the AHP system can heat the return water through the rectifier, condenser and absorber, and provide heating capacity to heat consumers. A higher PEE can be acquired resulting from additional low-grade heat utilization, including the ambient heat (either air-source or ground-source) utilized through the evaporator, and the flue gas waste heat recovered through the intermediate evaporator.

Fig. 1. The district heating network based on the proposed AHP system with intermediate process. 3

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the system high pressure (generation and condensation pressure) and low pressure (absorption and evaporation pressure), as is shown in Fig. 2(b). Compared to the single-effect AHP system shown in Fig. 2(a), the additional intermediate pressure can reduce the system performance degradation caused by large pressure difference under low ambient temperatures. Moreover, the introduction of intermediate process has two additional benefits: reducing waste heat discharge temperature and improving internal heat recovery. In conventional single-effect system, the waste heat of flue gas is either released directly to the ambient or recovered by the return water through a condensing heat exchanger. For the former, the waste heat discharge temperature is over 150 °C, and for the latter it is reduced to 60–70 °C, which is still higher than the dew

2.1. Principles of the intermediate process The main difference between the proposed system and the singleeffect system lies in the introduction of intermediate process, shown in the red dotted rectangle in Fig. 1. Flue gas of the combustor enters the intermediate evaporator, and its waste heat is recovered through intermediate evaporation process. Part of the ammonia liquid is evaporated and the two-phase flow enters the separator, where the liquid and vapor flow to the evaporator and the intermediate absorber, respectively. The vapor is absorbed by the weak solution from the generator, and the heat generated from the intermediate absorption process is recovered by the pressurized strong solution. The above mentioned processes are happened in an intermediate pressure, which is between

Fig. 2. System comparison based on P-T diagrams. 4

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point temperature of flue gas, resulting in latent heat not being utilized. Even worse, it raises the return water temperature and thus has a negative effect on the system performance, as will be shown in Section 4.3. In the proposed system these drawbacks can be eliminated, since the waste heat is recovered by the working fluid rather than circulating water. On the other hand, the conventional single-effect system usually uses a solution heat exchanger to achieve internal heat recovery. It transfers heat from the weak solution from the generator to the strong solution from the absorber. However, it usually has large temperature difference at the hot end resulting from the different mass flow rate of the two streams [44]. In the proposed system, the intermediate absorber is introduced to replace the solution heat exchanger, and it recovers not only the sensible heat of the solution, but also the heat released from the absorption of ammonia vapor. The extra intermediate absorption heat can reduce the temperature mismatch of the two streams and thus reduce the irreversible loss of the heat transfer process, as will be discussed in Section 4.2.

preheater and intermediate absorber, simulating the waste heat recovery process, which will be discussed in detail in Section 3.1. In the intermediate-process heating mode, the system operates as is described in Section 2.1. Besides, the generator can operates with or without internal heat recovery, and the rectifier can either be water-cooling or solutioncooling. Each is controlled by a set of valves, namely V13/V14 and V7/ V8. In this study, V14 and V8 are normally closed, and V13 and V7 are normally opened in the heating modes. That is, the generator operates with internal heat recovery, and the rectifier is water-cooling.

3. Methods of experimental study During the experimental study, the flue gas is simulated by the conduction oil, and principles of the simulation process is shown in this section. Besides, principles of calculation of key parameters and analysis of uncertainty are also clarified.

2.2. Experimental prototype system According to the previous theoretical and simulation studies [42], an experimental prototype based on the proposed AHP system with intermediate process is designed and built, as is shown in Fig. 3(a). The prototype consists of three units: the oil heating unit, the water cooling unit and the AHP unit. The oil heating unit mainly includes an electric oil heater, which simulates the natural gas combustor in Fig. 1. The oil heater heats the oil to the target temperature and delivers it to the generator of the AHP unit through the oil pipes; the water cooling unit mainly includes a water chiller, which simulates the heat consumer in Fig. 1. The water chiller cools the supply water down to the target return water temperature and delivers it to the AHP unit through the water pipes. Both units are electrically driven. As for the AHP unit, its structure is shown in Fig. 3(b), and its flow configuration and measuring points arrangement is shown in Fig. 4. The prototype has two heating modes including the single-effect mode and intermediate-process mode, as well as two auxiliary modes including charging mode and defrosting mode. The solution circulating valve V1 opens only in the charging mode, in which case valves V2, V3 and V4 are all closed. The absorbent (water) and the refrigerant (ammonia) is charged in sequence to the solution tank through valve V5. During this time, the solution pump drives the working fluid from the solution tank to the absorber and then back to the solution tank again, in order to transfer absorption heat to the circulating water. Otherwise the pressure of the solution tank will increase gradually and prevent further working fluid charging. In this mode, some refrigerant (ammonia) is also charged to the separator through valve V6, in order to reduce the starting time in the heating modes. After finishing charging, V1 closes, V3, V7 and V9 open, and part of the solution is pumped from the solution tank to the generator. The defrosting mode only works when the AHP unit is air-source and the evaporator is a finned tube or other similar forms. After a long period of operation at low ambient temperatures, the surface of the evaporator will be frosted, and the heat transfer process will deteriorate. At this time the defrosting mode starts, and the defrosting valve V10 opens. Part of the solution flows directly to the evaporator, where it absorbs the ammonia vapor and releases absorption heat to defrost. During the defrosting process, the system can still operate inefficiently, with COP and PEE lower than 1. In the single-effect heating mode, valve V11 closes and valve V12 opens. The solution preheater and intermediate evaporator do not transfer any heat. In this case, the separator changes into a tank with only ammonia liquid inside. Then valve V13 closes, and the intermediate absorber changes into a solution heat exchanger. As a result, the AHP unit becomes a single-effect system. In the intermediate-process heating mode, V11 and V12 open. Part of the heat conduction oil flows in sequence through the solution

Fig. 3. Pictures of the experimental prototype. 1-Generator, 2-Rectifier, 3Solution preheater, 4-Pressure transmitter, 5-Condenser, 6-Subcooler, 7-Mass flowmeter, 8-Intermediate evaporator, 9-Separater, 10-Temperature transmitter, 11-Evaporator, 12-Solution tank, 13-Content gauge, 14-Concentration meter, 15-Volume flowmeter, 16-Solution pump, 17-Absorber, 18-Intermediate absorber, 19- Electric control cabinet, 20-Display panel. 5

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Fig. 4. Flow configurations and measuring points arrangement of the AHP unit.

3.1. Simulation of waste heat recovery

3.2. Calculation of key parameters

In real applications, the flue gas of a natural gas combustor or a gasfired boiler is at a temperature of over 150 °C [33], containing a small amount of sensible heat and a large amount of latent heat. To fully recover such waste heat, the flue gas is first introduced to the solution preheater, where most of the sensible heat is recovered to heat the strong solution before entering the generator; then the flue gas enters the intermediate absorber, where the low-grade latent heat is recovered through the intermediate evaporation process. According to the previous study [42], its outlet temperature TWH,out, as well as the waste heat recovery ratio RWH, is influenced by the intermediate pressure PM, as is shown in Fig. 5 [42]. The waste heat ratio RWH is defined as the ratio of the recovered waste heat QWH to the generation heat QG, as is shown in Eq. (1):

(1) Calculation of the generation heat QG

RWH =

Q WH QG

The generation heat QG is defined as the enthalpy difference between heat conduction oil inlet and outlet of the generator. It is calculated based on Eq. (2):

QG = mOG ·c p,O·ΔTOG = ρO ·VOG·c p,O·(TOG,in − TOG,out )

(2)

where mOG and VOG are the mass and volume flow rates of the heat conduction oil through the generator; ρO and c p,O are the linear mean values of its density and specific heat, considering that they both change with temperature. The type of the heat conduction oil is Therminol 55, and its physical properties are shown in Fig. 7. (2) Calculation of the recovered waste heat QWH

(1)

In the intermediate-process heating mode of the experimental prototype, part of the heat conduction oil flows through the generator as the heating medium, with flow rate of VOG; the other part flows in sequence through the solution preheater and intermediate absorber, with flow rate of VOIE, and simulates the waste heat recovery process. During the experiment, the VOIE is carefully determined by adjusting valve V11 to achieve accurate simulation, and this process is based on the flow chart shown in Fig. 6, where TOG,in and TOG,out are the oil inlet and outlet temperatures of the generator, TWH,in and TWH,out are the waste heat inlet and outlet temperatures. The calculation of the generation heat QG and the recovered waste heat QWH will be introduced in the following Section 3.2. Besides, the heat transfer area of the solution preheater and intermediate absorber is carefully designed to ensure that the relative amount of heat exchanged in the two components satisfies the actual situation. More specifically, the ratio of the heat exchanged in the solution preheater to that in the intermediate absorber, under the intermediate pressure PM of 1 MPa, is around 0.2.

Fig. 5. Waste heat recovery under different intermediate pressure based on previous simulation study.[42] 6

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Fig. 7. Physical properties of Therminol 55 under different temperatures.

(3) Calculation of the heating capacity QHC In the experimental prototype, the return water passes in sequence through the condenser, rectifier and absorber, and are heated to the target temperature. Therefore, the heating capacity QHC consists of three parts: condensation heat QC, rectification heat QR and absorption heat QA. They are calculated based on Eqs. (6)–(9):

QC = mW ·c p,W ·ΔTWC = ρ W ·VW ·c p,W ·(TWC,out − TWC,in )

(6)

QR = mW ·c p,W ·ΔTWR = ρ W ·VW ·c p,W ·(TWR,out − TWR,in )

(7)

QA = mW ·c p,W ·ΔTWA = ρ W ·VW ·c p,W ·(TWA,out − TWA,in )

(8)

QHC = QC + QR + QA = ρ W ·VW ·c p,W ·(TSW − TRW )

(9)

where mW and VW are the mass and volume flow rates of the circulating water; ρ W and c p,W are its density and specific heat; TWC,in and TWC,out are the water inlet and outlet temperatures of the condenser; TWR,in and TWR,out are the water inlet and outlet temperatures of the rectifier; TWA,in and TWA,out are the water inlet and outlet temperatures of the absorber; TSW and TRW are the supply water and return water temperatures, TSW is equal to TWA,out, and TRW is equal to TWC,in. (4) Calculation of the heating efficiency COP and PEE The experimental prototype has two heating efficiency indexes: the coefficient of performance COP and the primary energy efficiency PEE, which are calculated based on Eqs. (10) and (11):

COP =

QHC QG

(10)

PEE =

QHC QHC ≈ QG/ ηC + WP/ ηE QG + Q WH + WP/ ηE

(11)

Fig. 6. Flow chart of the oil flow rate determination.

The recovered waste heat QWH consists of two parts: waste heat recovered in the solution preheater QSP and waste heat recovered in the intermediate evaporator QIE. They are calculated based on Eqs. (3)–(5):

QSP = mOIE ·c p,O·ΔTOSP = ρO ·VOIE·c p,O·(TOSP,in − TOSP,out )

(3)

Q IE = mOIE ·c p,O·ΔTOIE = ρO ·VOIE·c p,O·(TOIE,in − TOIE,out )

(4)

Q WH = QSP + Q IE = ρO ·VOIE·c p,O·(TWH,in − TWH,out )

(5)

where WP is the pump work; ηC is the efficiency of the combustor; ηE is the efficiency of electricity generation, which is around 33% [2]; QG/ ηC is the calorific value of the natural gas, and is approximately equal to the combustion heat plus flue gas waste heat: QG/ ηC ≈ QG + Q WH [42]. (5) Calculation of the internal heat recovery in the intermediate absorber QIA The intermediate absorber is a coaxial spiral double-pipe heat exchanger. In order to enhance heat and mass transfer, the pipe wall is processed into a corrugated shape, as is shown in Fig. 8. The intermediate absorption process happens in the outer pipe, where ammonia vapor is absorbed in the ammonia-water solution. The absorption heat is transferred through the corrugated pipe wall to the reverse-flowing cooling stream in the inner pipe, that is the strong solution from the

where mOIE and VOIE are the mass and volume flow rates of the heat conduction oil through the intermediate evaporator; TOSP,in and TOSP,out are the oil inlet and outlet temperatures of the solution preheater, and TOSP,in is equal to TWH,in; TOIE,in and TOIE,out are the oil inlet and outlet temperatures of the intermediate evaporator, and TOIE,out is equal to TWH,out. 7

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Fig. 8. Structures and thermometers arrangement of the intermediate absorber.

where ω T,x1, ω T,x2 , ω T,x3…,ω T,xn are the total uncertainties of x1, x2, x3,…, xn, respectively. For example, for the widely-used heat duty calculation equation:

solution pump. Thermometers are arranged at the inlet and outlet, as well as each trisection point of the heating and cooling streams. In this study, the calculation of the internal heat recovery QIA is based on the cooling stream (strong solution), considering that it is pure liquid and has no phase change during the heating process. QIA is defined as:

Q = ρ ·V ·c p·(Tin − Tout ) the total uncertainty of the heat duty ω T,Q can be calculated by:

Q IA = mSS ·c p,SS (x , P , T )·ΔTSSIA = ρSS (x , P , T )·VSS·c p,SS (x , P , T )·(TSSIA,out − TSSIA,in )

ω T,Q =

ω T,RWH =

ω T,PEE =

where ωS represents a range of possible dispersion of the measured value of a parameter to its truth value caused only by undetermined system error; ωR represents a range of possible dispersion caused only by the random error. For an indirectly-measured parameter, its total uncertainty ω T can be calculated by the related directly-measured parameters. For example, if an indirectly-measured parameter y is a function of directlymeasured parameters x1, x2, x3, …, xn:

2

2

2

⎛ ∂f ω T,x ⎞ + ⎛ ∂f ω T,x ⎞ + ⎛ ∂f ω T,x ⎞ + ···+⎛ ∂f ω T,x ⎞ 1 2 3 n ⎠ ⎠ ⎝ ∂x n ⎠ ⎝ ∂x2 ⎝ ∂x1 ⎝ ∂x3 ⎠ ⎟

⎜

⎟

⎜

)

(17)

QG2 (ωT,2QIE + ωT,2QSP ) + (Q IE + QSP )2ωT,2QG QG2

(18)

2 QG2 ωT,2QHC + QHC ωT,2QG

QG2

(19)

2 + QHC (ωT,2QG + ωT,2QIE + ωT,2QSP + ωT,2WP / ηE2 )

(QG + Q IE + QSP + WP/ ηE )2

(20)

The startup sequence of the experimental prototype follows this order: firstly the oil heating unit starts, then the water cooling unit starts, and finally the AHP unit starts. Before the AHP unit starts, the generator and rectifier are isolated from all other components, in order to quickly reach the required generation temperature. Fig. 9 shows the startup performance of the experimental prototype. It should be noted that a test run was conducted the day before, and the solution temperature in the generator reduced to about 60 °C after overnight cooling. An approximately linear pressure increase in the generator is observed after starting the oil heating unit, while a turning point in the solution temperature occurred 24 min after starting up, representing the solution’s transition from supercooled to saturated state. When the solution temperature and pressure reach 90% of the target

then the total uncertainty ω T,y can be calculated by: ⎜

⎟

4. Results and discussion

(14)

2

⎜

(QG + Q IE + QSP + WP/ ηE )2ωT,2QHC

(13)

⎟

⎟

While for the total uncertainty of waste heat ratio RWH, COP and PEE, the equations are:

For a directly-measured parameter, its total uncertainty ω T consists of two parts: the system uncertainty ωS and the random uncertainty ωR [10]:

⎜

⎜

(

3.3. Analysis of uncertainty

y = f (x1, x2 , x3, …, x n )

2 ⎛ ∂f ω T,V ⎞ + ⎛ ∂f ω T,Tin ⎞ + ⎛ ∂f ω T,Tout ⎞ ⎠ ⎝ ∂V ⎝ ∂Tin ⎠ ⎝ ∂Tout ⎠

= ρc p· (Tin − Tout )2ωT,2V + V 2 ωT,2Tin + ωT,2Tout

ω T,COP =

ω T = ωS + ωR

2

2

(12)

where mSS and VSS are the mass and volume flow rates of the strong solution; TSSIA,in and TSSIA,out are the strong solution inlet and outlet temperatures of the intermediate absorber; c p,SS (x , P , T ) and ρSS (x , P , T ) are the specific heat and density of the strong solution, and are functions of concentration x, pressure P and temperature T. The pressure and temperature can be easily acquired, and the concentration is measured by a concentration meter placed in the solution tank, as is shown in Fig. 3(b). After acquiring these parameters, c p,SS (x , P , T ) and ρSS (x , P , T ) can be calculated with the help of the software Aspen Plus.

ω T,y =

(16)

⎟

(15) 8

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return water temperatures is 34.24 °C. Detailed key parameters and the uncertainty analysis are shown in Table 1. The return water of 34.24 °C passes in sequence through the condenser, rectifier and absorber, and its temperature increases to 42.16, 46.36 and 55.09 °C, respectively, with flow rate keeping at 1.15 m3/h. It is shown that by introducing the intermediate process, the waste heat outlet temperature is reduced to 45.76 °C, which is lower than the dew point of the flue gas, indicating that the prototype has ability to recover the flue gas latent heat. Besides, after going through the internal heat recovery in the generator, the generated solution outlet temperature is reduced to 143.6 °C, and is 19.17 °C lower than the generation temperature. On the other hand, the pressurized strong solution of 36.29 °C is first heated to 130.40 °C by recovering the intermediate absorption heat; then it is further heated in the solution preheater, with its temperature remaining nearly constant, indicating that the solution has already changed to two-phase flow. It is also shown in Table 1 that the system high pressure, intermediate pressure and low pressure are 1.66 MPa, 1.03 MPa and 0.41 MPa, respectively. Moreover, the intermediate absorption pressure is 23.9 kPa lower than the intermediate evaporation pressure, and the absorption pressure is 14.3 kPa lower than the evaporation pressure, due to the flow resistance and the obstruction of check valves. Based on the measuring parameters, the heat duty or power consumption of each component, as well as the performance of the experimental prototype in the intermediate-process heating mode is shown in Table 2. Moreover, the uncertainty of each calculating parameter is calculated according to Eqs. (15)–(20). Table 2 shows that the energy input and output of the experimental prototype are 29.04 and 27.98 kW, respectively. Therefore, the heat loss is about 1.06 kW, which is only 4% of the total energy input, indicating that the experimental prototype is in good energy balance. The waste heat ratio RWH is

Fig. 9. Startup performance of the experimental prototype.

values, the water cooling unit starts. The generated ammonia vapor is condensed in the rectifier, and then flows into the solution. This process causes fluctuations in the solution temperature and pressure, but the overall trend is upward. 46 min after starting up, the pressure reaches 1.6 MPa, and the AHP unit starts and the experimental prototype begins to operate.

4.1. Performance under typical working condition The performance of the intermediate-process model is analyzed in detail under the typical working condition: the evaporation temperature is −0.58 °C, the generation temperature is 162.77 °C, and the

Table 1 Detailed key parameters and the uncertainty analysis under the typical working condition. Parameters

Average Value

System Uncertainty

Random Uncertainty

Total Uncertainty

Oil inlet temperature of the generator Oil outlet temperature of the generator Oil flow rate through the generator Oil inlet temperature of the solution preheater Oil outlet temperature of the solution preheater Oil outlet temperature of the intermediate evaporator Oil flow rate through the intermediate evaporator Water inlet temperature of the condenser Water outlet temperature of the condenser Water outlet temperature of the rectifier Water outlet temperature of the absorber Water flow rate Solution bottom temperature of the generator Solution inlet temperature of the intermediate absorber Solution upper temperature of the intermediate absorber Solution lower temperature of the intermediate absorber Solution outlet temperature of the intermediate absorber Solution recuperation inlet temperature of the intermediate absorber Solution recuperation lower temperature of the intermediate absorber Solution recuperation upper temperature of the intermediate absorber Solution recuperation outlet temperature of the intermediate absorber Solution recuperation outlet temperature of the solution preheater Refrigerant outlet temperature of the rectifier Refrigerant outlet temperature of the condenser Refrigerant outlet temperature of the subcooler Refrigerant outlet temperature of the expansion valve I Refrigerant outlet temperature of the intermediate evaporator Refrigerant outlet temperature of the expansion valve II Refrigerant outlet temperature of the evaporator Strong solution concentration Strong solution flow rate Generation pressure Intermediate evaporation pressure Intermediate absorption pressure Evaporation pressure Absorption pressure

188.75 °C 164.22 °C 1.32 m3/h 167.74 °C 147.41 °C 45.76 °C 0.043 m3/h 34.24 °C 42.16 °C 46.36 °C 55.09 °C 1.15 m3/h 162.77 °C 143.60 °C 85.44 °C 57.46 °C 46.29 °C 36.29 °C 53.70 °C 78.70 °C 130.40 °C 130.45 °C 75.34 °C 45.18 °C 37.80 °C 35.48 °C 45.18 °C −0.58 °C 1.73 °C 23.6% 0.15 m3/h 1657.8 kPa 1032.0 kPa 1008.1 kPa 423.1 kPa 408.8 kPa

± 0.10 °C ± 0.10 °C ± 0.01 m3/h ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.003 m3/h ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.01 m3/h ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.10 °C ± 0.3% ± 0.01 m3/h ± 5.0 kPa ± 5.0 kPa ± 5.0 kPa ± 5.0 kPa ± 5.0 kPa

± 0.28 °C ± 0.16 °C ± 0.01 m3/h ± 0.29 °C ± 0.25 °C ± 0.26 °C ± 0.002 m3/h ± 0.13 °C ± 0.12 °C ± 0.10 °C ± 0.16 °C ± 0.01 m3/h ± 0.49 °C ± 0.52 °C ± 0.22 °C ± 0.25 °C ± 0.36 °C ± 0.20 °C ± 0.23 °C ± 0.28 °C ± 0.55 °C ± 0.50 °C ± 0.46 °C ± 0.24 °C ± 0.19 °C ± 0.33 °C ± 0.28 °C ± 0.16 °C ± 0.28 °C ± 0.2% ± 0.01 m3/h ± 7.7 kPa ± 6.2 kPa ± 5.6 kPa ± 6.4 kPa ± 5.8 kPa

± 0.38 °C ± 0.26 °C ± 0.02 m3/h ± 0.39 °C ± 0.35 °C ± 0.36 °C ± 0.005 m3/h ± 0.23 °C ± 0.22 °C ± 0.20 °C ± 0.26 °C ± 0.02 m3/h ± 0.59 °C ± 0.62 °C ± 0.32 °C ± 0.35 °C ± 0.46 °C ± 0.30 °C ± 0.33 °C ± 0.38 °C ± 0.65 °C ± 0.60 °C ± 0.56 °C ± 0.34 °C ± 0.29 °C ± 0.43 °C ± 0.38 °C ± 0.26 °C ± 0.38 °C ± 0.5% ± 0.02 m3/h ± 12.7 kPa ± 11.2 kPa ± 10.6 kPa ± 11.4 kPa ± 10.8 kPa

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Table 2 Thermodynamic performance of the experimental prototype in intermediate-process mode. Energy Transfer

Performance Parameters

Components

Energy input (kW)

Energy output (kW)

Value

Uncertainty

Value

Uncertainty

Generator Rectifier Condenser Evaporator Intermediate Evaporator Absorber Solution Preheater Solution Pump Total

16.87 / / 8.76 2.17 / 0.45 0.79 29.04

± 0.41 / / ± 0.45 ± 0.25 / ± 0.05 ± 0.04 ± 0.66

/ 5.64 10.63 / / 11.71 / / 27.98

/ ± 0.41 ± 0.47 / / ± 0.48 / / ± 0.79

Parameters

Value

Uncertainty

Heating Capacity QHC Waste Heat Ratio RWH COP PEE

27.98 kW 0.16 1.66 1.28

± 0.79 kW ± 0.02 ± 0.06 ± 0.04

intermediate process, two main improvements are achieved. Firstly, the recovered heat increases from 12.8 to 19.5 kW, resulting from the absorption of the additional ammonia vapor. Secondly, the hot end temperature difference is reduced by more than 80%, demonstrating that the extra intermediate absorption heat has reduced the temperature mismatch of the two streams. Performance of the two heating modes under different evaporation temperatures is further studied. Fig. 12 shows that the strong solution concentration increases with the evaporation temperature TE, while the intermediate-process mode has a higher solution concentration than the single-effect mode, and the concentration difference becomes larger under higher TE. It is because part of the absorption process happens in the intermediate absorber, and this part is enhanced with a higher solution concentration resulting from a higher absorption pressure. Fig. 13 shows the heat duty of each component and the heating capacity in the two heating modes. The absorber produces more heat in the intermediate-process mode than in the single-effect mode as TE is −5 °C, but nearly the same as TE is 10 °C. Besides, more heat is recovered from the ambient in the intermediate-process mode, and the difference becomes more significant under lower TE. As a result, its heating capacity is larger than that of the single-effect mode, while the generation heat is nearly the same. Such an improvement can also be observed in Fig. 14, where the variations of COP in the two heating modes under different evaporation temperatures are presented. When TE is 15 °C, the COPs in both heating modes are about 1.95; when TE is 0 °C, the COP in the intermediate-process mode is 1.66, which is 6% larger than that in the single-effect mode; while when TE is −15 °C, the

0.16 as the intermediate pressure PM is 1.03 MPa, in line with the target value shown in Fig. 5. Moreover, the waste heat recovered in the solution preheater is about 21% of that in the intermediate evaporator, which is also consistent with the simulation value discussed in Section 3.1. It is demonstrated that when the evaporation temperature is −0.58 °C, the experimental prototype in the intermediate-process heating mode, can provide 28 kW heating capacity to heat the water from 34.24 to 55.09 °C, and the COP and PEE are 1.66 and 1.28, respectively. According to the previous simulation study [42], the COP is 1.85 under the similar working condition: the return and supply water temperatures are 35 and 55 °C, and the evaporation temperature is 0 °C. Therefore, the COP of the experimental prototype reaches nearly 90% of the simulation value, as is shown in Fig. 10, demonstrating that the prototype acquires good practical performance. Among the total 10% difference between the simulation and experiment values, 4% is mainly due to heat losses of the system, since the heat loss is 4% of the total energy input. The other 6% results from the flow resistances and pressure losses in pipes and heat exchangers, as well as heat transfer temperature differences. However, when the evaporation temperature reduces, the differences between the simulation and experiment values becomes larger, and the COP of the experimental value reaches 77% of the simulation value as the evaporation temperature reduces to −15 °C. It is probably because the heat and pressure losses will enlarge at low evaporation temperatures.

4.2. Comparison between two heating modes As is mentioned above, the experimental prototype has two heating modes: the single effect mode and the intermediate-process mode. In the single effect mode, the solution preheater and intermediate evaporator do not transfer any heat. In this case, the separator changes into a tank with only ammonia liquid inside. The intermediate absorber changes into a solution heat exchanger, and it transfers heat from the weak solution from the generator to the strong solution from the absorber. However, it usually has large hot end temperature difference, resulting from the different mass flow rate of the two streams. In the intermediate-process mode, the intermediate absorber recovers not only the sensible heat of the solution, but also the heat released from the absorption of the ammonia vapor. In order to study the difference of the heat transfer process in the intermediate absorber, T-Q diagram [45] of the two heating modes is drawn, as is shown in Fig. 11. The red and blue solid lines represent the heating and cooling streams in the intermediate-process mode, respectively, and the dashed lines represent those in the single-effect mode. It is concluded that by introducing the

Fig. 10. Comparison between the experimental results and previous simulation results [42]. 10

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Fig. 11. T-Q diagram of the intermediate absorber in the two heating modes.

Fig. 12. Strong solution concentration in the two heating modes. Fig. 13. Heat duty of each component and the heating capacity in the two heating modes.

COP in the intermediate-process mode is 1.28, which is almost 11% larger than that in the single-effect mode. That is, the COP enhancement ratio increases to 11% when the evaporation temperature reduces from 15 to −15 °C. Based on the above analysis, it can be obtained that the introduction of the intermediate evaporation and absorption can reduce the system performance degradation caused by large pressure difference under low ambient temperatures. It is proved that the proposed experimental prototype with intermediate process could be a more efficient way of residential district heating, especially in cold regions. 4.3. Influence of working condition parameters In Section 4.2, the influence of the evaporation temperature on the system performance is analyzed. In this section, many other working condition parameters in the intermediate-process heating mode are taken into consideration, including the generation temperature TG, intermediate pressure PM, the return water temperature TRW and the water flow rate VW. Their influences on the COP and supply water temperature TSW are analyzed separately when keeping other parameters as constants. The generation temperature TG is defined as the solution bottom temperature of the generator. It is an important working condition parameter as it influences the generation process and thus the system performance. A lower TG cannot provide sufficient driving force for efficient operating, while a higher TG may lead to larger irreversible loss during the heat transfer process. Many literatures indicate that for an absorption system, there exists an optimum TG, corresponding to the best system performance. When TG is higher than the optimum value, the performance will not improve or even reduce slightly. Such

Fig. 14. System performance the two heating modes under different evaporation temperatures.

phenomenon is also observed in this study, as is shown in Fig. 15. It can be concluded that when TG increases from 142 to 159 °C, the COP first increases rapidly and then decreases gradually. The turning point is around 150 °C, which is the optimum value when the evaporation temperature TE is 0 °C, the return water temperature TRW is 35 °C and the intermediate pressure PM is 1.1 MPa. Under the optimum condition, the COP of the experimental prototype is 1.67; while when TG gets larger, the COP is reduced to 1.65. For the supply water temperature TSW, it increases linearly from 54 to 59 °C when TG increases from 142 to 159 °C, and is expected to continue to raise when TG further 11

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Fig. 18. Influence of the water flow rate on the system performance.

Fig. 15. Influence of the generation temperature on the system performance.

other hand, a lower PM can enhance the waste heat recovery but hinder the absorption process. Therefore, there must be an optimal value of PM, corresponding to the best system performance. According to Fig. 16, as PM increases from 1 MPa to 1.28 MPa, the COP first raises from 1.62 to 1.69, and then reduces more and more slowly to 1.6. The maximum COP occurs when PM is about 1.1 MPa. When PM gets close to the system high pressure 1.66 MPa, the experimental prototype is equivalent to operating in the single-effect mode, and the COP is reduced to the corresponding single-effect value, which is about 1.55 shown in Fig. 14. The return water temperature TRW influences almost all the processes including the condensation, rectification, generation and absorption processes. A higher TRW makes the ammonia vapor more difficult to condense in the condenser and rectifier, and thus raises the system high pressure, which has a negative effect on the generation process. Even worse, it can also hinder the absorption process since the absorption heat is harder to carry away. In the previous analysis, TRW is stable at 35 °C, which is a common return water temperature for most district heating systems. However, it will change if the working condition deviates from the set state, such as the ambient temperature suddenly drops. Fig. 17 indicates that the system is quite sensitive to the return water temperature when the water flow rate VW is fixed. It is shown that the COP drops linearly from 1.52 to 1.4, as TRW increases from 33 to 37 °C. If TRW further increases to 43 °C, the COP reduces to only 1.2, while the supply water temperature TSW raises from 55 to 58 °C. On the other hand, if the return water temperature TRW is fixed, the COP will increase and the supply water temperature TSW will decrease, as the water flow rate VW raises, as is shown in Fig. 18. Therefore, if TRW deviates from the set value, VW can be adjusted accordingly to stabilize the system performance. More specifically, enlarge VW if TRW increases, and reduce VW if TRW decreases.

Fig. 16. Influence of the intermediate pressure on the system performance.

5. Conclusions Fig. 17. Influence of the return water temperature on the system performance.

In this paper, an ammonia-water absorption heat pump prototype with intermediate process for residential district heating is designed and built. The prototype is driven by an electric oil heater, which simulates the natural gas combustor, and heat conduction oil is used as the heating medium. To perform comparison study, it has two heating modes: the single-effect mode and intermediate-process mode. Experimental investigations are carried out to analyze the system performance under the typical working condition, and comparison between the two heating modes is performed. Moreover, influences of various internal and external parameters are further analyzed, including the evaporation temperature, generation temperature, intermediate pressure, return water temperature and water flow rate, in order to analyze the system performance under different working

increases. Therefore, in order to obtain the targeted supply water temperature, the generation temperature can be appropriately increased above the optimal value, at the cost of slightly reducing the COP. The intermediate pressure PM influences the intermediate evaporation and absorption processes, as well as the waste heat recovery. Based on the previous study, a higher PM is beneficial to the intermediate absorption process but leads to higher intermediate evaporation temperature, which reduces the waste heat recovery. This phenomenon is demonstrated in Fig. 16, since the recovered waste heat QWH decreases from 2.91 to 2.35 kW as PM increases from 1 MPa to 1.28 MPa. On the

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interests or personal relationships that could have appeared to influence the work reported in this paper.

conditions. Several conclusions are drawn as the following: (1) When the evaporation temperature is about 0 °C, the experimental prototype in the intermediate-process heating mode, can provide 30 kW heating capacity to heat the water from 34.24 to 55.09 °C, and the coefficient of performance COP and primary energy efficiency PEE are 1.66 and 1.28, respectively. When the evaporation temperatures reduce to −5, −10 and −15 °C, the COPs are 1.51, 1.40 and 1.28, respectively. (2) The COP of the experimental prototype reaches nearly 90% of the simulation value when the evaporation temperature is 0 °C. Among the total 10% difference, 4% is mainly due to heat losses of the system, since the heat loss is 4% of the total energy input. The other 6% results from the flow resistances and pressure losses in pipes and heat exchangers, as well as heat transfer temperature differences. However, the COP of the experimental value reaches 77% of the simulation value as the evaporation temperature reduces to −15 °C. (3) Compared to the single-effect system, the system with intermediate process performs better at lower evaporation temperature, with an improvement of 11% in COP. Moreover, the recovered heat in the intermediate absorber increases from 12.8 to 19.5 kW, and the hot end temperature difference is reduced by more than 80%. (4) The COP first increases rapidly and then decreases gradually, and the supply water temperature increases linearly, as the generation temperature increases from 142 to 159 °C, and its optimal value is about 150 °C. In order to obtain the targeted supply water temperature, the generation temperature can be appropriately increased above the optimal value, at the cost of slightly reducing the COP. (5) The recovered waste heat decreases by nearly 20% as the intermediate pressure PM increases from 1 MPa to 1.28 MPa, while at this time the COP first raises, and then reduces more and more slowly. The maximum COP occurs when PM is about 1.1 MPa. When PM gets close to the system high pressure, the experimental prototype is equivalent to operating in the single-effect mode, and the COP is reduced to the corresponding single-effect value. (6) The COP drops and the supply water temperature TSW raises gradually, as the return water temperature TRW increases and the water flow rate VW keeps constant. On the other hand, if the return water temperature TRW is fixed, the COP increases and the TSW decreases as the VW raises. Therefore, if the TRW deviates from the set value, VW can be adjusted accordingly to stabilize the system performance. More specifically, enlarge VW if TRW increases, and reduce VW if TRW decreases.

Acknowledgements The authors would like to thank the support of the National Science Fund for Distinguished Young Scholars (No. 51625603) and the International Partnership Program of Chinese Academy of Sciences (No. GJHZ1876). References [1] Tsinghua University Building Energy Saving Research Center. Annual Report on China Building Energy Efficiency. Beijing: China Architecture and Building Press; 2013. [in Chinese]. [2] Yang L, Yan H, Lam JC. Thermal comfort and building energy consumption implications – a review. Appl Energy 2014;115(Supplement C):164–73. [3] Tsinghua University Building Energy Saving Research Center. Annual Report on China Building Energy Efficiency. Beijing: China Architecture and Building Press; 2017. [in Chinese]. [4] Li J. Towards a low-carbon future in China's building sector—a review of energy and climate models forecast. Energy Policy 2008;36(5):1736–47. [5] Li W, Lin X, Cao C, et al. Organic Rankine Cycle-assisted ground source heat pump combisystem for space heating in cold regions. Energy Convers Manage 2018;165:195–205. [6] Keinath CM, Garimella S, Garrabrant MA. Modeling of an ammonia–water AHP water heater for residential applications. Int J Refrig 2017;83:39–50. [7] Dai E, Lin M, Xia J, et al. Experimental investigation on a GAX based absorption heat pump driven by hybrid liquefied petroleum gas and solar energy. Sol Energy 2018;169:167–78. [8] Wu W, Shi W, Li X, et al. Air source absorption heat pump in district heating: applicability analysis and improvement options. Energy Convers Manage 2015;96:197–207. [9] Han BC, Cheng WL, Nian YL, et al. Evaluation of an ammonia-water absorption long-distance cooling/heating system for enhanced geothermal resources. Energy Convers Manage 2019;192:346–58. [10] Wu W, Shi W, Wang J, et al. Experimental investigation on NH3-H2O compressionassisted absorption heat pump (CAHP) for low temperature heating under lower driving sources. Appl Energy 2016;176:258–71. [11] Qunli Z, Mingkai C, Qiuyue Z, et al. Research on a new district heating method combined with hot water driven ground source AHP. Energy Procedia 2015;75:1242–8. [12] Zhang Q, Zhang X, Sun D, et al. Municipal space heating using a ground source absorption heat pump driven by an urban heating system. Geothermics 2019;78:224–32. [13] Wu W, Ran S, Shi W, et al. NH3-H2O water source AHP (WSAHP) for low temperature heating: experimental investigation on the off-design performance. Energy 2016;115:697–710. [14] Nitkiewicz A, Sekret R. Comparison of LCA results of low temperature heat plant using electric heat pump, absorption heat pump and gas-fired boiler. Energy Convers Manage 2014;87:647–52. [15] Li Y, Fu L, Zhang S, et al. A new type of district heating system based on distributed absorption heat pumps. Energy 2011;36(7):4570–6. [16] Garrabrant M, Stout R, Glanville P, et al. Development and validation of a gas-fired residential heat pump water heater-final report. Stone Mountain Technologies, Inc.; 2013. [17] Baig AA, Ethirveerasingham K, Fung A. Development of a monitoring system for a gas-fired absorption heat pump and results for testing in heating and cooling mode. Energy Procedia 2019;158:2202–9. [18] Garimella S, Christensen RN, Lacy D. Performance evaluation of a generator-absorber heat-exchange heat pump. Appl Therm Eng 1996;16(7):591–604. [19] Chatzitakis P, Safarov J, Opferkuch F, et al. Experimental investigation of an absorption heat pump with organic working pairs. Appl Therm Eng 2019;163:114311. [20] Wu W, Wang B, Shi W, et al. An overview of ammonia-based absorption chillers and heat pumps. Renew Sustain Energy Rev 2014;31:681–707. [21] Cai D, Jiang J, He G, et al. Experimental evaluation on thermal performance of an air-cooled absorption refrigeration cycle with NH3–LiNO3 and NH3–NaSCN refrigerant solutions. Energy Convers Manage 2016;120:32–43. [22] Abdulateef JM, Sopian K, Alghoul MA. Optimum design for solar absorption refrigeration systems and comparison of the performances using ammonia-water, ammonia-lithium nitrate and ammonia-sodium thiocyanate solutions. Int J Mech Mater Eng 2008;3(1):17–24. [23] Táboas F, Bourouis M, Vallès M. Analysis of ammonia/water and ammonia/salt mixture absorption cycles for refrigeration purposes in fishing ships. Appl Therm Eng 2014;66(1–2):603–11. [24] Farshi LG, Infante Ferreira CA, Mahmoudi SMS, Rosen MA. First and second law analysis of ammonia/salt absorption refrigeration systems. Int J Refrig 2014;40:111–21. [25] Fukuta M, Yanagisawa T, Iwata H, et al. Performance of compression/absorption hybrid refrigeration cycle with propane/mineral oil combination. Int J Refrig 2002;25(7):907–15. [26] Wu W, Wang B, Shang S, et al. Experimental investigation on NH3–H2O

By replacing the original heat exchangers, the proposed system can be easily applied to the existing heat networks based on gas-fired boilers to improve energy efficiency, and the favorable experimental performance proves that it is a more efficient way of residential district heating, especially in cold regions. CRediT authorship contribution statement Ding Lu: Conceptualization, Methodology, Formal analysis, Investigation, Writing - original draft, Software, Visualization. Yin Bai: Investigation, Validation, Data curation. Yanxing Zhao: Writing - review & editing, Funding acquisition. Xueqiang Dong: Writing - review & editing, Software. Maoqiong Gong: Supervision, Project administration, Funding acquisition. Ercang Luo: Project administration. Gaofei Chen: Visualization. Qingyu Xu: Validation. Jun Shen: Resources. Declaration of Competing Interest The authors declare that they have no known competing financial 13

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[27] [28] [29]

[30] [31] [32]

[33]

[34]

[35]

[36] Che D, Liu Y, Gao C. Evaluation of retrofitting a conventional natural gas fired boiler into a condensing boiler. Energy Convers Manage 2004;45(20):3251–66. [37] Yang B, Jiang Y, Fu L, et al. Experimental and theoretical investigation of a novel full-open absorption heat pump applied to district heating by recovering waste heat of flue gas. Energy Build 2018;173:45–57. [38] Zhang H, Zhao H, Li Z. Waste heat recovery and water-saving modification for a water-cooled gas-steam combined cycle cogeneration system with absorption heat pump. Energy Convers Manage 2019;180:1129–38. [39] Qu M, Abdelaziz O, Yin H. New configurations of a heat recovery AHP integrated with a natural gas boiler for boiler efficiency improvement. Energy Convers Manage 2014;87:175–84. [40] Zhu K, Xia J, Xie X, et al. Total heat recovery of gas boiler by AHP and directcontact heat exchanger. Appl Therm Eng 2014;71(1):213–8. [41] Wei M, Yuan W, Song Z, et al. Simulation of a heat pump system for total heat recovery from flue gas. Appl Therm Eng 2015;86:326–32. [42] Lu D, Chen G, Gong M, et al. Thermodynamic and economic analysis of a gas-fired absorption heat pump for district heating with cascade recovery of flue gas waste heat. Energy Convers Manage 2019;185:87–100. [43] Gong M, Lu D, Chen G, Zhao Y. A gas-fired absorption heat pump system. China Patent; 201811277686.3. [44] Lu D, Xu Q, Chen G, et al. Modeling and analysis of an ammonia–water absorption refrigeration system utilizing waste heat with large temperature span. Int J Refrig 2019;103:180–90. [45] Du S, Wang RZ, Xia ZZ. Graphical analysis on internal heat recovery of a single stage ammonia–water absorption refrigeration system. Energy 2015;80:687–94.

compression-assisted absorption heat pump (CAHP) for low temperature heating in colder conditions. Int J Refrig 2016;67:109–24. Liu C, Wang Z, Han W, et al. Working domains of a hybrid absorption-compression heat pump for industrial applications. Energy Convers Manage 2019;195:226–35. Wu W, You T, Wang J, et al. A novel internally hybrid absorption-compression heat pump for performance improvement. Energy Convers Manage 2018;168:237–51. Liu X, Ye Z, Bai L, et al. Performance comparison of two absorption-compression hybrid refrigeration systems using R1234yf/ionic liquid as working pair. Energy Convers Manage 2019;181:319–30. Xu ZY, Wang RZ. Absorption refrigeration cycles: categorized based on the cycle construction. Int J Refrig 2016;62:114–36. Venegas M, Izquierdo M, De Vega M, et al. Thermodynamic study of multistage absorption cycles using low-temperature heat. Int J Energy Res 2002;26(8):775–91. Li X, Wu W, Zhang X, et al. Energy saving potential of low temperature hot water system based on air source absorption heat pump. Appl Therm Eng 2012;48:317–24. Shang S, Li X, Chen W, et al. A total heat recovery system between the flue gas and oxidizing air of a gas-fired boiler using a non-contact total heat exchanger. Appl Energy 2017;207:613–23. Li F, Duanmu L, Fu L, et al. Research and application of flue gas waste heat recovery in co-generation based on absorption heat-exchange. Procedia Eng 2016;146:594–603. Li Y, An H, Li W, et al. Thermodynamic, energy consumption and economic analyses of the novel cogeneration heating system based on condensed waste heat recovery. Energy Convers Manage 2018;177:671–81.

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