Performance optimization of heat pump driven liquid desiccant dehumidification systems

Energy and Buildings 52 (2012) 132–144

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Performance optimization of heat pump driven liquid desiccant dehumidification systems Tao Zhang, Xiaohua Liu ∗ , Yi Jiang Department of Building Science, Tsinghua University, Beijing 100084, China

a r t i c l e

i n f o

Article history: Received 21 December 2011 Received in revised form 27 May 2012 Accepted 9 June 2012 Keywords: Liquid desiccant Dehumidification Regeneration Performance optimization Air-conditioning

a b s t r a c t Heat pump driven liquid desiccant dehumidification (HPLD) devices, in which both the cooling capacity from an evaporator and heat from a condenser are utilized, have been developing rapidly in recent years. Because the amount of heat from the condenser is usually more than required for desiccant regeneration, the key to improve the performance of these systems involves effectively exhausting the extra heat. In this paper, two different methods for removing the extra heat are analyzed: adding an air-cooled assistant condenser in either the inlet or outlet of the regeneration air duct, and utilizing a water-cooled assistant condenser that uses cooled water from evaporation of the exhaust air. These handling processes were compared in terms of their HPLD performance under various operating conditions. Devices using an aircooled or water-cooled assistant condenser showed better performance than a basic system without an assistant condenser. For systems with an air-cooled assistant condenser, the location of the condenser rarely affected the performance: COPsys was an average of about 18% higher than that of the basic system. The system with the water-cooled condenser performed the best out of all the investigated systems: COPsys was about 35% higher than that of the basic system. © 2012 Elsevier B.V. All rights reserved.

1. Introduction The major tasks of air-conditioning systems are to control both the indoor temperature and the humidity. The dehumidification of air with the help of liquid desiccant has been developing rapidly in recent years. The differences between air and liquid desiccant in terms of their heat and mass transfer performance have been widely investigated by scholars both theoretically and experimentally [1–5]. Based on both theoretical models and experimental results, various air handling processors have been designed and some of them have been applied in commercial buildings [6–9]. Heat pump driven liquid desiccant (HPLD) systems have become more and more popular recently due to their compact size and high efficiency. Both the cooling capacity from an evaporator and heat from a condenser are utilized in these systems; the cooling capacity is used to cool the desiccant to enhance its dehumidification ability and the heat is used to regenerate the desiccant. For HPLD systems, there are different ways to utilize the condensing heat for desiccant regeneration. Yadav [10], Ma et al. [11] and Yamaguchi et al. [12] put all the condensing heat from the heat pump in the solution for regeneration in their HPLD systems. In studies conducted by Kinsara et al. [13] and Dai et al. [14], all the

∗ Corresponding author. Tel.: +86 10 6277 3772; fax: +86 10 6277 0544. E-mail address: [email protected] (X. Liu). 0378-7788/$ – see front matter © 2012 Elsevier B.V. All rights reserved. http://dx.doi.org/10.1016/j.enbuild.2012.06.002

condensing heat from the heat pump was removed by the regeneration air rather than by the solution, and the heated air went into the regenerator for desiccant regeneration. In a heat pump cycle, the condensing heat Qc equals sum of the compressor power and the cooling capacity Qe . However, as the amount of condensing heat is usually greater than the heat that is actually needed for desiccant regeneration [15], some scholars have suggested adding water into the regenerator to remove the extra condensing heat, including Zhao et al. [8] and Zhu et al. [9]. Water is added to the regenerator to lower the condensing temperature of the heat pump, but contradiction does exit in the regenerator, aiming at producing concentrated desiccant but mixing the desiccant with water. In the HPLD system designed by Lazzarin and Castellotti [16], only part of the condensing heat from the condenser was utilized to regenerate the desiccant, and the extra heat was used to preheat the inlet regeneration air in the regenerator. In order to match the evaporating and condensing capacities, Niu et al. [15] and Zhang et al. [17] suggested an auxiliary air-cooled condenser after the solutioncooled condenser to extract the extra condensing heat of a HPLD system. HPLD systems use the condensing heat from the heat pump in different ways. However, to obtain better mass transfer performance, previous research has recommended that the regeneration heat be used to heat the inlet solution rather than the inlet air in the regenerator [18]. As a result, in the systems evaluated in this paper, the desiccant is heated by the condenser before coming into

T. Zhang et al. / Energy and Buildings 52 (2012) 132–144

133

Nomenclature a COPhp COPhp,id COPsys cp h Le ˙ m NTUm Pcom Ppump Qc Qe r t V

specific area of packed-bed per volume (m2 /m3 ) COP of the heat pump COP of the ideal Carnot heap pump COP of the dehumidification process taking heat recovery into consideration specific heat capacity (kJ/kg ◦ C) enthalpy (kJ/kg) Lewis number (no unit of measurement) mass flow rate (kg/s) number of mass transfer units between the air and desiccant power consumption of the compressor (kW) power consumption of the circulating pumps in the HPLD system (kW) condensing heat of the heat pump (kW) cooling capacity of the heat pump (kW) latent heat of vaporization (kJ/kg) temperature (◦ C) volume of heat and mass transfer module (m3 )

Greek symbols ˛ heat transfer coefficient between air and desiccant (kW/m2 ◦ C) ˛m mass transfer coefficient between air and desiccant (kg/m2 s)  concentration (mass ratio of desiccant to solution) of liquid desiccant (%) ε thermodynamical perfectness of the heat pump humidity ratio (g/kg) ω Subscripts a air cond condensation dehumidifier deh e air in equilibrium with desiccant ea exhaust air evap evaporation fresh air fa in inlet out outlet r refrigerant return air ra reg regenerator s liquid desiccant supply air sa

the regenerator; only the extra condensing heat that is not required for desiccant regeneration is removed by the assistant condenser. Three HPLD systems using different types of assistant condensers were compared to a basic HPLD system without an assistant condenser. Several available models, such as a heat and mass transfer model, a heat pump model, and a heat exchanger model, were adopted to simulate the performance of the HPLD systems. Based on these validated models, the different systems were compared in terms of their performance. It is hoped that the results of this research will be helpful for the optimization and application of these types of liquid desiccant systems. 2. Different HPLD systems According to mass conservation, the moisture removal rate of processed air will be equal to that removed by regenerated air.

Fig. 1. The Basic Type HPLD system.

Therefore, in the HPLD system, the latent heat of the processed air side is the same as that of the regenerated air side. However, in the heat pump cycle, the condensing heat is always greater than the cooling capacity of an evaporator (about 20–25% greater when COP is 4–5). If the entire amount of condensing heat is used to regenerate the desiccant, the desiccant temperature in the condenser will increase, leading to a further increase of the condensing temperature of the heat pump and a decreased COP of the heat pump. Thus, in HPLD systems, improving system performance involves using the condensing heat efficiently for desiccant regeneration and effectively discharging the extra condensing heat. The Basic Type HPLD system is shown in Fig. 1, and the operating principles of the three different HPLD systems are shown in Fig. 2. In these processes, the cooling capacity of the evaporator is utilized to cool liquid desiccant coming into the dehumidifier. A small part of the diluted solution after dehumidification is sent into the regenerator and the same amount of concentrated solution after regeneration returns back to the dehumidifier. There are two heat recovery devices: one is used for the heat recovery between the diluted solution sent to the regenerator and the concentrated solution sent to the dehumidifier, and the other is used for the total heat recovery of the indoor return air. In the Basic Type HPLD system shown in Fig. 1, fresh air comes into the heat recovery device to recover the energy from the indoor return air, and then it goes into the desiccant-sprayed dehumidifier to be further dehumidified and cooled. The return air goes into the regenerator after heat recovery, and it removes the moisture from the diluted solution, thus achieving desiccant regeneration. All the condensing heat from the heat pump is put into the solution for regeneration in the Basic Type system, which is similar to the system proposed by Yadav [10]. In the Type I system shown in Fig. 2(a), an air-cooled assistant condenser is added in the HPLD system. After heat recovery, the return air goes into an air-cooled condenser to remove part of the condensing heat, and then the heated air goes into the regenerator and serves as the regenerating air. This process is similar to the system introduced by Lazzarin and Castellotti [16]. In this process, part of the condensing heat is utilized to heat the inlet regeneration air, and most of the condensing heat is put into the solution for regeneration. The major difference between the Type I and Type II systems is the location of the air-cooled assistant condenser. In the Type II system, after heat recovery, the return air first goes into the regenerator for regeneration and then flows through the air-cooled condenser to take away the extra condensing heat. In the Type III system shown in Fig. 2(c), a water-cooled assistant condenser is utilized instead of the air-cooled condenser in the

134

T. Zhang et al. / Energy and Buildings 52 (2012) 132–144

Fig. 2. Different HPLD systems: (a) Type I with an air-cooled assistant condenser (b) Type II with an air-cooled assistant condenser and (c) Type III with a water-cooled assistant condenser.

Type I and II systems. After heat recovery, the return air goes into the regenerator for desiccant regeneration and then flows through an evaporative cooler. Because of the heat and mass transfer process between the return air and the cycling water, the outlet water from the evaporating cooler can be used as the cooling water for the water-cooled condenser. In this process, part of the condensing heat is removed by the water-cooled condenser, and most of the condensing heat is used for desiccant regeneration. The main differences between the Type I–III systems can be summarized as follows: the Type I system uses an air-cooled assistant regenerator located at the inlet of the regeneration air duct; the Type II system uses an air-cooled assistant regenerator located

at the outlet of the regeneration air duct; and the Type III system uses a water-cooled assistant condenser with cooled water from the evaporation of the exhaust regeneration air. 3. Theoretical models of components 3.1. Heat and mass transfer model Packed-bed towers are common handling devices for liquid desiccants, and the dehumidification module of a three-dimensional cross-flow packed-bed tower is shown in Fig. 3(a). The coupled heat and mass transfer process between the air and the

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135

Fig. 3. Heat and mass transfer model of cross-flow between air and liquid desiccant: (a) three-dimensional schematic and (b) two-dimensional schematic.

solution proceeds both in dehumidification and regeneration; the x-dimension is the liquid desiccant flow direction and the z-dimension is the air flow direction. Fig. 3(b) shows a twodimensional model that considers the symmetrical distribution of air and solution. This heat and mass transfer process can be decoupled, as proposed by Liu et al. [19]. Energy conservation, mass conservation between air and solution, and solute mass conservation are shown in Eqs. (1)–(3), respectively. ˙ s hs ) = m ˙ s · hs + m ˙ s · dhs = 0 ˙ a · dha + dm ˙ a · dha + d (m m

(1)

˙ s=0 ˙ a · dωa + dm m

(2)

˙ s · ) = 0 d(m

(3)

The heat and mass transfer equations are shown in Eqs. (4)–(6).



NTUm · Le dha = · (he − ha ) + r · V dV dωa NTUm = · (ωe − ωa ) V dV Le =

˛ , ˛m cp,e

NTUm =

1

Le





− 1 (ωe − ωa )

(4) (5)

˛m aV ˙a m

(6)

To describe the heat and mass transfer between air and water, energy conservation and mass conservation can be expressed by Eqs. (1) and (2), respectively, by replacing the parameter of liquid desiccant with the parameter of water; the corresponding heat and mass transfer equations can be expressed by Eqs. (4)–(6) in a similar way. 3.2. Heat pump system model Fig. 4 shows the refrigerant cycle of a heat pump in the HPLD system. Point 2 to point 3 shows the compression process in the compressor; point 6 to point 2 shows the evaporation process; and point 3 to point 5 shows the condensation process. In this cycle, it is assumed that the condensing pressure and the evaporating pressure of the refrigerant remain constant. Point 5 to point 6 shows the inflation process in throttle; the refrigerant enthalpies before and after inflation are equal. Compression efficiency, compression volume efficiency, and motor efficiency were compared with the parameters of an ideal compressor model to simulate the performance of the actual compressor. Point 2 is the refrigerant inlet state of the compressor shown in Fig. 4, and outlet refrigerant enthalpy can be calculated by Eq. (7): h3 =

h3

id

− h2

c

+ h2

where h3 id and h3 are the ideal and practical refrigerant enthalpies of the outlet refrigerant, respectively; h2 is the inlet refrigerant enthalpy of the compressor, which is the same as the outlet refrigerant enthalpy of the evaporator; and c is the compression efficiency of the compressor and varies with the ratio of condensing and evaporating pressure of the refrigeration cycle [20]. ˙ r can be calculated by Eq. (8): Refrigerant flow rate m ˙r= m

Vs

v2

v

(8)

where Vs is the standard compressor volume, 2 is the inlet refrigerant specific volume of the compressor, and v is the compression volume efficiency. Then, energy consumption of the compressor Wc can be calculated by Eq. (9): Wc =

˙ r (h3 − h2 ) m m

(9)

where m is the motor efficiency. Inlet refrigerant enthalpy h5 in throttle is equal to outlet enthalpy h6 . Assuming the refrigerant pressure in the condensation and evaporation processes remains constant, condensing heat Qc and cooling capacity Qe can be calculated by Eqs. (10) and (11), respectively: ˙ r · (h3 − h5 ) Qc = m

(10)

˙ r · (h2 − h6 ) Qe = m

(11)

In the numerical simulations, the cold temperature in the condenser (subcooling) and the hot temperature in the evaporator

P 5

3_id

4

Pc

P 6

3

1

2

h

(7) Fig. 4. P–h diagram of the refrigerant cycle of the heat pump.

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T. Zhang et al. / Energy and Buildings 52 (2012) 132–144

Fig. 5. Flow chart of the calculating method for simulating the performance of a HPLD system.

(superheating) are given as 3 ◦ C and 5 ◦ C, respectively. COPhp of the heat pump can be calculated by: COPhp =

Qe Pcom

(12)

According to the second thermodynamical law, the COP of ideal Carnot heap pump operating at condensing temperature tcond and evaporating temperature tevap is shown in Eq. (13). Therefore, the thermodynamical perfectness of the actual heat pump to the ideal Carnot heat pump ε can be obtained by Eq. (14). COPhp,id = ε=

tevap + 273.15 tcond − tevap

COPhp COPhp,id

(13)

(14)

3.3. Models of other components 3.3.1. Heat exchanger model There are various kinds of heat exchangers in different HPLD systems. The condenser and the evaporator of the heat pump are the heat exchangers between the refrigerant and the solution. The heat recovery module between the return air and the fresh air is an air-to-air total heat exchanger, and the heat recovery module between the diluted solution and the concentrated solution is a liquid-to-liquid sensible heat exchanger. The ε-NTU method was used to describe the performance of heat exchangers. Heat transfer capacity is the product of the heat transfer coefficient and the heat transfer area, and the total heat transfer capacities of the condensers in the different HPLD systems examined in this study are identical. For the Type I–III systems shown in Fig. 2, it is necessary to distribute the heat transfer capacity between

T. Zhang et al. / Energy and Buildings 52 (2012) 132–144

137

18 15 12 9 6 3 0

Test Simulation 1 stage 2 stages

Power / kW

5 4

1 stage

2 stages 3 stages

3

Test

2

Simulation

1 4

1

(c)

3 stages

(b)

7 10 Data point

13

(d)

31

4

1

16

7 10 Data point

4

21

3 stages

Simulation 1 stage

11

2 stages

13

16

Test

Test COPhp

tsa /

(a)

ωsa / (g/kg)

Fig. 6. Operating schematic of a three-stage HPLD system.

3

Simulation 3 stages

2

1 stage

2 stages

1

1 1

4

7 10 Data point

13

1

16

4

7 10 Data point

13

16

Fig. 7. Test data and simulated results for validation of the theoretical models from Zhang et al. [17]: (a) supply air temperature, (b) supply air humidity ratio, (c) power consumed by compressors and (d) COP of the heat pump.

Table 1 Tested and simulated results of a three-stage HPLD system by Zhang et al. [21]. Mode

tfa (◦ C)

One stage

26.24 26.24 26.63 20.97 22.06 21.46 20.07

8 9 10 11

Two stages

12 13 14 15 16

Three stages

Data point

1 2 3 4 5 6 7

tea (◦ C)

ωea (g/kg)

Tested

Simulated

Tested

Simulated

5.08 5.02 4.87 6.57 6.51 6.21 6.30

33.90 35.50 36.60 34.60 35.56 34.10 33.10

14.23 14.91 15.35 13.40 14.14 13.88 13.56

12.00 12.40 12.83 12.50 11.90 11.40 11.10

11.74 12.01 12.13 10.06 10.32 10.08 9.32

3.52 4.29 4.49 3.72 4.37 4.00 3.80

3.80 4.13 4.10 4.00 4.28 4.22 4.03

8.12 7.68 7.78 8.27 7.52 7.55 7.74

7.39 7.74 7.76 7.39 7.43 7.39 7.25

2.83 2.50 2.32 2.38 2.63 2.61 2.49

2.66 2.63 2.58 2.52 2.91 2.74 2.64

27.43 27.43 27.93 28.03

10.78 10.94 10.72 10.67

36.90 38.90 40.50 40.00

17.06 23.14 24.38 23.96

7.20 7.40 7.40 7.20

7.67 7.77 7.84 7.83

3.67 3.74 3.79 3.76

4.03 3.97 3.97 3.96

17.63 17.40 17.48 17.50

17.45 17.48 17.54 17.54

2.31 2.34 2.32 2.33

2.21 2.25 2.24 2.24

30.51 30.61 30.21 30.02 29.72

12.02 11.85 12.34 13.44 13.70

39.20 39.00 41.60 42.50 42.20

23.48 24.06 24.62 32.67 32.19

1.60 1.80 2.10 2.30 2.60

1.79 1.78 2.08 2.62 2.68

2.78 2.82 2.87 2.92 3.00

2.69 2.68 2.75 2.88 2.89

25.70 26.09 26.49 26.58 26.58

26.26 26.28 26.52 26.81 26.85

1.86 1.81 1.79 1.86 1.86

1.86 1.82 1.80 1.86 1.87

ωfa (g/kg)

tsa (◦ C)

ωsa (g/kg)

Power (kW)

COPhp

Tested

Tested

Simulated

Simulated

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T. Zhang et al. / Energy and Buildings 52 (2012) 132–144

Table 2 Operating performance data for the four HPLD systems under different operating conditions. tfa (◦ C)

Operating condition

Condition A

Condition B

Condition C

Condition D

Condition E

Condition F

Basic Type Type I Type II Type III Basic Type Type I Type II Type III Basic Type Type I Type II Type III Basic Type Type I Type II Type III Basic Type Type I Type II Type III Basic Type Type I Type II Type III

36

ωfa (g/kg)

22

36

18

36

14

32

22

32

18

32

14

tsa (◦ C) 24.82 22.54 20.70 19.07 25.55 23.92 22.26 20.27 26.64 25.60 24.31 22.09 23.08 20.89 19.23 18.47 23.80 22.03 20.57 19.24 24.82 23.64 22.48 20.61

ωsa g/kg

9.60

9.60

the air-cooled (Types I and II) or the water-cooled (Type III) assistant condenser and the solution-cooled condenser. As the purpose of the auxiliary condenser is to extract the excessive condensing heat [15,17], the distribution should take into consideration both the heat required for desiccant regeneration and an optimal condensing heat discharge. To simplify analysis, 20% of the heat transfer capacity of the condenser in the Type I–III systems is distributed to the assistant condenser; this amount was used in the subsequent simulations of the different HPLD systems in present research. However, different distribution ratios of the heat transfer capacity affects the performance of HPLD systems and the detailed analysis will be investigated in further research.

tea (◦ C)

ωea (g/kg)

tevap (◦ C)

tcond (◦ C)

COPsys

ε

48.00 47.97 50.44 35.13 43.70 42.76 44.95 33.28 38.32 38.49 40.15 31.11

25.08 25.01 24.84 30.68 20.75 20.96 20.90 26.06 16.90 17.00 16.96 21.43

8.85 4.17 3.67 5.07 11.88 9.94 8.49 7.76 17.07 15.48 13.77 11.45

70.60 57.16 54.96 46.46 61.03 49.10 48.16 42.01 48.96 42.64 42.26 37.15

1.91 2.42 2.54 3.37 2.70 3.64 3.57 4.27 4.69 5.68 5.36 6.00

3.42 3.83 3.92 4.92 4.13 5.29 5.08 5.69 6.13 7.10 6.65 6.81

0.42 0.46 0.47 0.50 0.46 0.50 0.50 0.52 0.52 0.53 0.53 0.54

45.04 45.08 46.73 33.47 40.73 40.34 42.12 31.70 35.93 36.38 37.83 29.76

25.12 24.90 24.83 29.41 20.78 20.96 20.91 25.12 16.91 16.99 16.96 20.83

8.30 4.56 4.58 6.86 11.33 9.40 8.41 8.94 16.04 14.35 12.95 11.43

66.95 53.36 50.89 43.96 57.37 46.21 45.21 39.82 46.37 40.28 39.90 35.54

2.16 2.72 2.92 3.87 2.95 3.92 3.92 4.84 4.99 5.98 5.71 6.45

3.43 4.06 4.26 5.46 4.21 5.31 5.22 6.08 5.83 6.69 6.38 6.64

0.45 0.48 0.49 0.51 0.48 0.51 0.51 0.53 0.52 0.54 0.54 0.55

COPhp

3.3.2. Mixing module model The mixing processes between the regenerated solution and the dehumidifying solution and between the diluted solution after dehumidification and the regenerating solution are shown in Fig. 1 (Basic Type) and Fig. 2 (Types I–III). The mixing process between solutions of different concentrations should follow the solute mass balance equation, the solution mass balance equation, and the energy balance equation, expressed by Eqs. (15)–(17), respectively: ˙ s,2 · 2 = m ˙ s,3 · 3 ˙ s,1 · 1 + m m

(15)

˙ s,2 = m ˙ s,3 ˙ s,1 + m m

(16)

Table 3 Solution states for the four HPLD systems under different operating conditions. Operating condition

Dehumidifier

Regenerator

ts,in (◦ C)

 s,in (%)

ts,out (◦ C)

 s,out (%)

ts,in (◦ C)

 s,in (%)

ts,out (◦ C)

 s,out (%)

Condition A

Basic Type Type I Type II Type III

15.76 12.49 11.42 11.50

47.51 45.29 41.10 35.33

24.43 21.78 19.99 18.84

47.27 45.03 40.89 35.17

57.81 50.92 46.74 40.88

49.01 46.79 42.32 36.22

44.23 44.89 38.63 35.48

49.29 47.08 42.55 36.39

Condition B

Basic Type Type I Type II Type III

18.03 16.09 14.46 13.26

47.01 44.26 41.26 36.60

25.19 23.17 21.36 19.54

46.80 44.07 41.10 36.48

51.19 44.68 42.19 37.67

48.21 45.31 42.18 37.32

40.99 40.42 36.20 33.48

48.43 45.51 42.35 37.46

Condition C

Basic Type Type I Type II Type III

21.27 19.81 18.17 15.85

45.91 44.42 42.49 38.66

26.14 24.80 23.26 20.90

45.77 44.29 42.37 38.57

43.09 39.71 38.15 34.19

46.66 45.12 43.14 39.20

37.05 36.88 34.03 31.34

46.80 45.25 43.26 39.30

Condition D

Basic Type Type I Type II Type III

14.81 12.04 11.33 12.46

45.84 43.04 38.45 33.10

22.98 20.65 19.09 18.76

45.60 42.81 38.26 32.96

55.03 47.62 43.81 39.07

47.24 44.38 39.51 33.86

42.40 42.10 36.77 34.42

47.52 44.63 39.71 34.00

Condition E

Basic Type Type I Type II Type III

17.09 15.10 13.86 13.70

45.34 42.08 38.97 33.99

23.75 21.63 20.11 19.05

45.14 41.91 38.83 33.89

48.41 42.04 39.83 35.96

46.44 43.04 39.79 34.60

39.16 38.06 34.51 32.30

46.65 43.21 39.94 34.72

Condition F

Basic Type Type I Type II Type III

19.99 18.42 17.04 15.37

44.35 42.51 40.54 36.50

24.54 23.07 21.74 19.86

44.23 42.40 40.44 36.41

40.92 37.48 36.10 32.74

45.06 43.16 41.13 36.97

35.33 34.80 32.33 30.07

45.19 43.28 41.24 37.06

T. Zhang et al. / Energy and Buildings 52 (2012) 132–144

139

Fig. 8. Heat analysis of the HPLD systems: (a) Basic Type and (b) Type II.

˙ s,2 · hs,2 = m ˙ s,3 · hs,3 ˙ s,1 · hs,1 + m m

(17)

where subscripts 1 and 2 express the two solutions before mixing and subscript 3 depicts the solution after mixing. On basis of the theoretical models, COPsys of the HPLD system is defined as the cooling capacity of the fresh air (Qfa ) divided by the power consumed by the compressor and the circulating pumps. COPsys =

Qfa

(18)

Pcom + Ppump

COPhp are 8%, 7.7%, 2.3% and 5.7%, respectively. This means that the variations of the measured data and the simulated results were consistent, and that the simulated results based on the theoretical models showed good agreement with the measured results. Thus, the theoretical models of the various components can be used to accurately analyze the performance of the different HPLD systems. 4. Performance comparison of the different HPLD systems 4.1. Energy performance in different operating conditions

3.4. Validation of the theoretical models Based on the theoretical models of the components, HPLD system model can be established on the Simulink platform in Matlab software. Choosing the Basic Type in Fig. 1 as an example, the flow chart of calculating method is shown in Fig. 5. Following the work of Zhang et al. [21], the performance of a three-stage HPLD system was measured and the differences between the test data and the simulated results were analyzed. The operating principle of the three-stage HPLD system is shown in Fig. 6 and all the condensing heat was put into the solution in this process. The system can adjust the number of operating stages according to the load ratio. A model of this HPLD system was built based on the theoretical models of the previously mentioned components to analyze the performance. The test data and the simulated results are shown in Fig. 7, including supply air temperature, humidity ratio, compressor power and COPhp . The detailed tested results and simulated results of this three-stage HPLD system are listed in Table 1. An index of the mean relative bias between the measured and simulated results was used to evaluate the validation of the theoretical models, expressed by Eq. (19): =

n   i  1

n



,

i =

Di − Di Di

× 100%

(19)

where is the mean relative bias between the measured and simulated results;  i is the bias of a certain data point; Di is the measured  data and Di is the simulated result of a certain data point; and n is the number of data points. As shown in Fig. 7 and listed in Table 1, the mean relative biases of supply air temperature, humidity ratio, compressor power and

Based on the mathematical models of the main components described in the previous section, system models were built for the HPLD systems shown in Figs. 1 and 2, and the differences in energy performance of these processes were investigated. Numerical simulations were carried out on the Simulink software platform. COPhp and COPsys were adopted to compare the performances of various HPLD systems. The performance data for the HPLD systems under typical operating conditions is listed in Table 2, with a lithium bromide (LiBr) aqueous solution used as the liquid desiccant. Table 3 gives the solution states of the four HPLD systems under different operating conditions. The parameters of return air were fixed as tra = 26 ◦ C and ωra = 12.6 g/kg, and the required humidity ratio of supplied air ωsa was fixed at 9.6 g/kg. The flow rates of fresh air and return air are both 4000 m3 /h. The mass flow rate of spray solute in the dehumidifier or the regenerator is 0.75 kg/s, and circulating solute between the dehumidifier and the regenerator is 0.1 kg/s. As the energy consumption of a circulating pump is much lower than the compressor in most cases, power consumed by a circulating pump is treated as constant to simplify the calculation. For the dehumidifier, regenerator and evaporative cooler modules, the power consumed by each circulating pump is regarded as 0.25 kW in the following analysis. v is given as 0.95 and m is given as 0.8. Choosing Condition A as an example, Fig. 8(a) and (b) gives the heat exchange rates of key components in Basic Type and Type II respectively. For these two HPLD systems, cooling capacity of the evaporator or condensing heat of the condenser equals to the sum of the heat and cold offset in the solution heat exchanger and the enthalpy variance of the air flowing through. As to the Basic

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Fig. 9. Operating performance of different processes in typical operating conditions: (a) supply air temperature, (b) COPsys , (c) evaporating temperature and (d) condensing temperature.

Type, all the condensing heat has to be removed by the solutioncooled condenser and heat removed by the air flowing through the regenerator is 50.3 kW, much larger than the cooling capacity (34.6 kW) extracted by the outdoor air flowing through the dehumidifier. While in Type II, part of the condensing heat (16.3 kW) is extracted by the auxiliary air-cooled condenser and the condensing heat removed by the air flowing through the regenerator is 36.7 kW, close to the cooling capacity (40.4 kW) extracted by the outdoor air. It indicates that in Type II the auxiliary condenser helps to remove the condensing heat efficiently and it is more balanced between the heat for regeneration and the condensing heat extracted by the solution-cooled condenser compared with the Basic Type. Fig. 9 shows the performance data for the various HPLD systems under the outdoor parameters of Conditions A and C listed in Table 2. Taking Condition A (tfa = 36 ◦ C, ωfa = 22 g/kg) as an example, the temperature of supplied air (tsa ) was highest in the Basic Type system (24.8 ◦ C) and lowest in the Type III system (19.1 ◦ C). Temperature difference T between the condensing temperature and the evaporating temperature of the heat pump was 61.8 ◦ C, 53.0 ◦ C, 51.3 ◦ C, and 41.4 ◦ C for the Basic Type, Type I, Type II, and Type III systems, respectively. A higher T resulted in a lower COP

of the heat pump and therefore a lower COPsys . Fig. 10 shows a psychrometric chart of the air handling processes of the four HPLD systems in Condition A. In the Basic Type system, the entire heat of the condenser is put into the liquid desiccant for regeneration. As the amount of condensing heat is greater than the actual amount required for desiccant regeneration, the desiccant has to be heated further. Therefore, the desiccant leaves the condenser with a much higher temperature, which led to a much higher condensing temperature than that in the Type I, II, or III system. In Condition A, the condensing temperature (tcond ) of the Basic Type system was as high as 70 ◦ C, and both its COPhp and COPsys were the lowest of all the HPLD systems. However, tevap of the Basic Type system was higher than that of the others. In the Type I and II systems, the condensing heat is partly put into the liquid desiccant for regeneration, and the extra heat is removed by the regeneration air (return air) directly. For these two systems, tcond decreased to about 55 ◦ C in Condition A, much lower than that of the Basic Type system. The two types of HPLD systems with aircooled assistant condensers showed similar performance: COPsys was an average of about 18% higher than that of the Basic Type.

T. Zhang et al. / Energy and Buildings 52 (2012) 132–144

141

Fig. 10. Psychrometric chart showing the air handling processes in a typical condition: (a) Basic Type, (b) Type I, (c) Type II and (d) Type III.

In the Type III system, part of the condensing heat is utilized for desiccant regeneration and the extra heat is exhausted by the evaporative cooler. This water-cooled assistant condenser can effectively exhaust the extra condensing heat and decrease the condensing temperature. In Condition A, tcond of the Type III system was lower than 47 ◦ C. Although the evaporating temperature was not the highest among the investigated HPLD systems, COPsys of the Type III system was the highest. In other words, the Type III system with the water-cooled assistant condenser removed the extra condensing heat more efficiently compared to the Type I or II system with the air-cooled assistant condensers. COPsys of the Type III system was about 35% higher than that of the Basic Type system. Detailed operating performance data for the HPLD systems under various operating conditions is shown in Table 2. As the supplied air had a constant humidity ratio requirement (9.6 g/kg), COPsys increased with the decrease of outdoor air temperature or humidity ratio. Effective extraction of the extra condensing heat will improve the energy performance of the systems, and evaporative cooling of the regeneration air (Type III) performed better than using the regeneration air to extract the heat directly (Types I and II). Thermodynamical perfectness ε of the heat pump is listed in Table 2, which is around 0.5 and for different working conditions.

4.2. Operating parameters of liquid desiccant Among the different HPLD processes, there were obvious differences in the operating parameters of liquid desiccant. Table 3 lists the temperatures and concentrations of inlet and outlet solutions for dehumidifier and regenerator under various operating conditions. The concentrations of liquid desiccant in the HPLD systems are shown in Fig. 11. The desiccant concentration is based on the balancing results of the system, the requirement of dehumidification, and how much condensing heat needs to be removed from the heat pump. To handle the processed air to the same humidity ratio in the dehumidifier, a lower balanced desiccant concentration means a lower desiccant temperature is required. And a lower desiccant concentration means it is easier to be concentrated by the regeneration air. It can be seen from Fig. 11 that in all operating conditions, the desiccant concentration of the Basic Type system was the highest and that of the Type III system with the watercooled assistant condenser was the lowest out of all investigated processes. Due to the fact that the amount of exhaust heat from the condenser is more than what is actually required for desiccant regeneration, the regenerating temperature of the Basic Type system has to be increased to remove the entire amount of condensing

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(a)

50

(b) 50

Basic Type

Basic Type

46

46

Type I 42

Type I

38

Type II

ξ /%

ξ /%

42 Type II 38 Type III 34

34 Type III

30

30

A

B

C

D

E

F

A

Operating condition

B

C

D

E

F

Operating condition

Fig. 11. Operating concentration of liquid desiccant in different operating conditions: (a) inlet concentration of the dehumidifier and (b) outlet concentration of the regenerator.

heat; this led to a much higher desiccant concentration. In the Type I and II systems, part of the condensing heat is removed by the return air directly, which helps to lower the condensing temperature; this resulted in relatively lower desiccant concentration levels for these processes. The inlet temperature of the regeneration air (return air side) into the regenerator in the Type I system was much higher than that in the Type II system because the air-cooled assistant condenser in the Type I system is located close to the inlet of the regeneration air duct. The desiccant concentration of the Type III system was the lowest out of all the systems due to its effective way to remove the extra condensing heat through evaporative cooling of the exhaust regeneration air. To summarize, the operating desiccant concentration decreased with the decrease of the condensing temperature of the heat pump. The variance between the balanced desiccant concentration and the outdoor conditions was quite different for the different HPLD processes, as indicated in Fig. 11. For the Basic Type system, the balanced desiccant concentration increased with the increase of the outdoor temperature or humidity ratio. For the Type III system, the desiccant concentration increased with the increase of the outdoor temperature or the decrease of the outdoor humidity ratio. For the

(a)

7.5

4.3. Energy performance analysis of the Type III system As discussed in the previous section, the Type III system with the water-cooled assistant condenser showed the best performance of all the investigated processes. This section summarizes the results of a more detailed analysis of the performance of the Type III system. The effects of outdoor air temperature and humidity ratio on the performance of the Type III system are shown in Fig. 12, where the required humidity ratio of the supplied air was 9.6 g/kg. It can be seen from the figure that COPhp increased with the decrease of the outdoor air temperature or the humidity ratio. COPhp decreased from 6.4 to 3.9 when outdoor air tfa was fixed at 32 ◦ C and ωfa varied from 14 g/kg to 22 g/kg. The seasonal performance of the Type III HPLD system can be calculated as long as the outdoor air parameters are known. Fig. 13 illustrates the statistical frequencies of different outdoor temperature and humidity ratio intervals of Beijing and Shanghai during air-conditioning season [22]. Based on the single-point energy

(b)

t =28

7.0

Type I and II systems, the effect of outdoor temperature on balanced desiccant concentration was not obvious or consistent.

7.0 t =28 6.5

6.5

t fa =32

5.5

COPsys

COP hp

6.0

5.0

6.0

5.5

t =32

4.5 4.0

5.0

t =36

t =36

3.5 3.0

4.5

14

16

18 20 ω fa (g/kg)

22

14

16

18 20 ω fa (g/kg)

22

Fig. 12. Energy performance of the Type III system for different outdoor parameters: (a) COP of the heat pump and (b) COP of the process.

T. Zhang et al. / Energy and Buildings 52 (2012) 132–144

(b) 38

(a) 38 36

1

1

10

10

42

85

79

60

109

34

46 27

31

32

t fa /

t fa /

12

36

32 21

32 30

30 49

26 12

4

4

5

34

28

143

66

79

55

17

14

17 22 ω fa /(g/kg)

28

27

26 12

17

156 87

22 ω fa /(g/kg)

27

Fig. 13. Statistical frequencies of outdoor hourly parameters during air-conditioning season: (a) Beijing and (b) Shanghai.

performance of the Type III system and relevant statistics, seasonal COPhp was calculated to be 5.4 in Beijing and 4.8 in Shanghai, while seasonal COPsys was calculated to be 6.2 in Beijing and 6.0 in Shanghai.

5. Conclusions Heat pump driven liquid desiccant dehumidification (HPLD) devices have been developing rapidly in recent years. Various HPLD systems can be developed because there are different ways to utilize the condensing heat for desiccant regeneration. The key to improving system performance involves discharging the extra condensing heat effectively. Three HPLD systems using different types of assistant condensers were compared with the Basic Type HPLD system without an assistant condenser. The results might be helpful for the optimization and future application of HPLD systems. Based on the results of the comparisons, the following conclusions can be made: (1) The HPLD systems that used air-cooled or water-cooled assistant condensers showed better energy performance compared to the Basic Type system that put all the condensing heat into the solution. (2) The location of the air-cooled assistant condenser (i.e., before or after the solution-cooled condenser) rarely affected the system performance according to comparisons between the Type I and Type II systems under different operating conditions. COPsys was an average of about 18% higher than that of the Basic Type system. (3) The Type III system that used cooling water from an evaporative cooler to discharge the extra condensing heat performed the best out of all the investigated HPLD systems. COPsys was about 35% higher than that of the Basic Type system, and seasonal COPsys of the Type III system was calculated to be 6.2 and 6.0 in Beijing and Shanghai, respectively.

Acknowledgements The research described in this paper was supported by National Natural Science Foundation of China (No. 51138005) and the foundation for the author of National Excellent Doctoral Dissertation of China.

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