refrigerant mixture as a working fluid

Energy 44 (2012) 1005e1016

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Thermodynamic analysis of an absorption refrigeration system with ionic-liquid/ refrigerant mixture as a working fluid Yoon Jo Kim a, *, Sarah Kim b, Yogendra K. Joshi c, Andrei G. Fedorov c, Paul A. Kohl b a

Department of Mechanical Engineering, Washington State University Vancouver , 14204 NE Salmon Creek Avenue, Vancouver, WA 98686-9600, USA School of Chemical and Biomolecular Engineering, Georgia Institute of Technology, Atlanta, GA 30332, USA c The George Woodruff School of Mechanical Engineering, Georgia Institute of Technology, Atlanta, GA 30332, USA b

a r t i c l e i n f o

a b s t r a c t

Article history: Received 6 December 2011 Received in revised form 17 April 2012 Accepted 20 April 2012 Available online 26 May 2012

Thermodynamics of an ionic-liquid (IL) based absorption refrigeration system has been numerically analyzed. It provides an alternative to the normally toxic working fluids, such as the ammonia in conventional absorption systems. The use of ILs also eliminates crystallization and metal-compatibility problems of the water/LiBr system. Mixtures of refrigerants and imidazolium-based ILs are theoretically explored as the working fluid pairs in a miniature absorption refrigeration system, so as to utilize waste-heat to power a refrigeration/heat pump system for electronics cooling. A non-random two-liquid (NRTL) model was built and used to predict the solubility of the mixtures. Saturation temperatures at the evaporator and condenser were set at 25
C and 50
C, respectively, with the power dissipation of 100 W. Water in combination with [emim][BF4] (1-ethyl-3-methylimidazolium tetrafluoroborate) gave the highest coefficient of performance (COP) around 0.9. The refrigerant/IL compatibility indicated by the circulation ratio, alkyl chain length of the IL, and thermodynamic properties of the refrigerants, such as latent heat of evaporation were proven to be important factors in determining the performance of the absorption system. The negative effect of high viscosity was mitigated by dilution of the IL with the refrigerant and the use of slightly larger microfluidic channel heat exchangers. Ó 2012 Elsevier Ltd. All rights reserved.

Keywords: Ionic liquid Absorption system Waste-heat Electronics cooling

1. Introduction Recent advances in semiconductor technologies have led to an increase in power density for high performance chips, such as microprocessors. According to the International Technology Roadmap for Semiconductors (ITRS), these chips are expected to dissipate an average heat flux as high as 75 W/cm2, with the maximum junction temperature not exceeding 85
C, in 2012, while in 2024 the numbers are more challenging, 120 W/cm2 and 70
C, respectively [1]. Conventional chip packaging solutions, which use aircooling, face difficulties in dissipating such high heat fluxes in the limited space allocated to thermal management. A variety of novel thermal solutions for electronic cooling have been reported, including thermosyphons [2], loop heat pipes [3], electro-osmotic pumping [4], stacked micro-channels [5], impinging jets [6], thermoelectric micro-coolers [7], vapor compression refrigeration [8], and absorption based refrigeration systems [9,10]. These cooling systems can be categorized into

* Corresponding author. Tel.: þ1 360 546 9184; fax: þ1 360 546 9438. E-mail address: [email protected] (Y.J. Kim). 0360-5442/$ e see front matter Ó 2012 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2012.04.048

passive and active systems. Passive cooling systems utilize capillary or gravitational buoyancy forces to circulate the working fluid, while active cooling systems are driven by a pump or compressor for higher cooling capacity and improved performance. Also, an active system may offer further increases in power dissipation by insertion of a negative thermal resistance into heat flow path [8]. In this study, ionic-liquids (ILs), which are salts in a liquid state usually with organic cations and inorganic anions, are used as an absorbent fluid in a miniature absorption refrigeration system designed for current electronic cooling requirements (i.e., benchmark 100 W/cm2 power dissipation and 85
C chip temperature). ILs have the character of molten salts, which are moisture and air stable at room temperature. Most ILs are thermally stable to temperatures well above those in vapor compression refrigeration systems, >400 K [11e15]. However, in the case of a prolonged exposure to elevated temperatures, the effective decomposition temperature could be lower. Blake et al. [15] reported that the halflife of [bmim][PF6] is only 138 days at 573 K, while at 423 K (the highest operating temperature of an absorption system), it could be more than 10 years. Since the degradation products of ILs are often volatile compounds [12], the operating temperature needs to

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Y.J. Kim et al. / Energy 44 (2012) 1005e1016

be kept within acceptable limits. Due to the toxic and flammable nature of volatile organic compounds (VOCs) [16], ILs with the negligible vapor pressure have been considered as a possible replacements of solvents for organic synthesis [17], biphasic catalysis, separation and extraction processes [18], and dissolution of biomaterials [19]. Note that although ILs are ‘non-flammable’, some of ILs could be ‘combustible’ at sufficiently high temperature (near their decomposition temperature) upon exposure to fire or combustion [20]. Also, Jastorff et al. [21] found the toxicity of the various ILs spanned a range of about 1000  (from least to most toxic) based on the toxicological tests of a series of ILs with imi   dazolium cations and [PF 6 ], [BF4 ], [Cl ], [Br ] and tosylate anions. However, even the most toxic of the ILs has about the same toxicity as the least toxic of four common organic solvents (methanol, acetone, acetonitrile, MTBE) that were tested using the same protocol [13]. Thus, even though some ILs may not be intrinsically “green”, they can be designed to be environmentally benign, with large potential benefits for sustainable chemistry [22]. While the chemistry of ILs and their utilization for chemical processes have been reported in many publications, solubility data with other solvents and thermodynamic properties of mixtures containing ILs and refrigerants are less common [11]. With the above mentioned features of ILs such as tunable properties, zero vapor pressure, and high thermal stability, ILs are promising absorbents [23]. In particular, the low volatility of the IL enables easy separation of the volatile working fluid from the IL by thermal stratification with the minimum harmful impacts on environment [24]. ILs can provide an alternative to the normally toxic working fluids used in some absorption systems, such as the ammonia/water system. Since many of the ILs have melting points below the lowest solution temperature in the absorption system (w300 K) [25e29], they also eliminate the crystallization and metal-compatibility problems of the water/LiBr system. Among the ILs considered in this study, the melting points of [emim][BF4] (1ethyl-3-methylimidazolium tetrafluoroborate) and [bmim][PF6] (1-butyl-3-methylimidazolium hexafluorophosphate) are 14.42
C [28] and 10
C [29], respectively, which are low enough to be used in absorption system. Several theoretical works on the absorption refrigeration system using ILs as absorbent have been reported [23,24,30e32], and an experimental system was investigated by Kim et al. [24]. In this study, the thermodynamic performance of a miniature absorption system using refrigerant/IL mixtures as the working fluid pairs was theoretically investigated. Various mixtures of refrigerants and imidazolium ILs e [emim][Tf2N] (1-ethyl3-methylimidazoliumbis(trifluoromethanesulfonyl)imide), [emim] [BF4](1-ethyl-3-methylimidazoliumtetrafluoroborate), [bmim][BF4] (1-butyl-3-methylimidazoliumtetrafluoroborate), [bmim][PF6] (1-butyl3-methylimidazoliumhexafluorophosphate), [hmim][Tf2N] (1-hexyl3-methylimidazoliumbis(trifluoromethanesulfonyl)imide), [hmim] [BF4] (1-hexyl-3-methylimidazoliumtetrafluoroborate), and [hmim] [PF6] (1-hexyl-3-methylimidazoliumhexafluorophosphate) e were compared among each other in terms of system efficiency and usability of low-grade waste heat as the heat source of the system. Several HFC refrigerants, R125 (pentafluoroethane), R134a (1,1,1,2-tetrafluoroethane), R143a (1,1,1-trifluoroethane), R152a (1,1-difluoroethane), and R32 (difluoromethane), were investigated as the working fluid in the IL-based refrigeration system. Relevant characteristics of refrigerants, including ozone depletion potentials (ODP) and global warming potentials (GWP) are given in Table 1. The performance of an IL with R114 (dichlorotetrafluoroethane) and R124 (chlorotetrafluoroethane) were evaluated and compared to hydrofluorocarbon refrigerants. Water, which is attractive due to its high thermal conductivity and latent heat of evaporation, was also explored as a refrigerant with IL absorbents.

Table 1 Saturation pressures and latent heats of various refrigerants at the evaporator and condenser temperatures (Te ¼ 25
C, Tc ¼ 50
C) and their ozone depletion potentials and global warming potentials [33]. Refrigerants Psat,e [kPa] Psat,c[kPa] Dhlv,e [kJ/kg] Dhlv,c [kJ/kg] ODPa GWPb R114 R124 R125 R134a R143a R152a R32 Water

214.4 382.7 1376.7 665.3 1262.3 596.4 1689.6 3.170

446.9 775.8 2533.2 1317.9 2307.9 1177.4 3141.2 12.35

128.07 146.56 110.4 177.8 159.6 279.4 270.9 2441.7

117.53 130.56 76.4 151.8 118.9 245.4 209.6 2382.0

1 0.02 0 0 0 0 0 0

3.9 620 3400 1300 4300 120 620 e

a

The ODP of all other refrigerants are compared to R11 [30]. GWP is a relative scale which compares the greenhouse gas to carbon dioxide where GWP by definition is 1 [30]. b

In all calculations, the operating temperatures of the condenser and evaporator were set at 50
C and 25
C, respectively. The cooling capacity of the evaporator was chosen to be 100 W. The degree of superheat at the evaporator outlet and subcooling at the condenser outlet were set at 5
C, with respect to saturation temperatures at the evaporator and condenser pressures. The outlet temperatures of the desorber and absorber were treated as adjustable parameters, which can be tuned to find the optimum operating conditions where the system efficiency reached its maximum value. 2. Waste-heat driven absorption system Fig. 1 shows a schematic diagram of a thermally driven absorption refrigeration system using a refrigerant/IL mixture as a working fluid pair. The system consists of an evaporator, an absorber, a desorber, a condenser, a liquid pump, and an expansion device. One of the major advantages of the absorption refrigeration system is the utilization of waste heat, which can have temperature of 90
C or lower. The use of waste-heat results in a significant reduction in operating cost and justifies the added balance-of-plant for the absorption refrigeration system. Furthermore, the only

Heat absorption from waste-heat(Qd)

Heat rejection to ambient (Qc)

Chemical compressor

3 Condenser Refrigerant 4 Expansion device 1

Desorber 8

Strong-IL Solution

7 Solution HX

6 Liquid pump 5

9 Expansion device 10

Evaporator

Weak-IL Solution

Absorber 2

Heat absorption from source (Qe)

Heat rejection to ambient (Qa)

Fig. 1. Schematic diagram of an absorption refrigeration system using IL/refrigerant mixture as a working fluid pair.

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

component of this system with moving mechanical parts is the liquid pump, so that relatively quiet operation is possible with no lubrication needed. In general, the key system performance metric, coefficient of performance (COP), is defined as the heat removed at the evaporator per total power supplied to the system, Equation (1).

Qe  COP ¼
Wp þ Qd

(1)

where Wp and Q are pumping work and the heat input/output, respectively. The subscripts “e” and “d” denote the evaporator and desorber, respectively. Since the heat supply at the desorber is from waste heat (i.e., it is essentially free), a more relevant coefficient of performance (COPwaste) can be defined by eliminating Qd from Equation (1), i.e., COPwaste ¼ Qe =Wp . The cyclic process for the refrigerant loop is the same as that of a vapor compression system, except the mechanical compressor is replaced with a ‘chemical compressor’ which consists of an absorber, liquid pump, solution heat exchanger, desorber and expansion device. The pressurization process in the chemical compressor starts in the absorber, where the refrigerant vapor from the evaporator (state point 2) is exothermically absorbed into the weak (refrigerant) solution (state point 10), resulting in strong (refrigerant) solution at state point 5. Once the refrigerant is absorbed, the solution is pressurized by the liquid pump. The solution heat exchanger preheats the strong solution of state point 6 to state point 7 using the high temperature weak solution flow from the desorber. A high pressure and high temperature superheated refrigerant vapor is generated in the desorber and the refrigerant is endothermically desorbed from the strong solution. The refrigerant vapor returns to the refrigerant loop. Meanwhile, the mixture solution becomes the weak solution and returns to the absorber through solution heat exchanger and expansion device in sequence, which completes the solution loop or chemical compression cycle. Table 2 summarizes the main components and their processes in the waste-heat driven absorption refrigeration system. The condensation/absorption process at the absorber and vaporization/desorption process at the desorber make it possible to use a low-power-input liquid pump to increase the pressure between the condenser and the evaporator. Although the presence of the absorber and desorber increases the overall system volume, the displacement volume and power consumption for compression of the liquid are much smaller than those for vapor compression. Also, the slight modification of the absorption refrigeration system can be used for power cycles, e.g., the Kalina cycle [34e36] and a heat transformer [37].

model by Shiflett and Yokozeki [38] are used in this study, Equations (2) and (3).

" lngm ¼ x2n snm " lngn ¼

x2m

smn



Fnm xm þ xn Fnm



Fmn xn þ xm Fmn

2 þ

snm Fnm

(2)

ðxn þ xm Fmn Þ2

2 þ

#

smn Fmn

# (3)

ðxm þ xn Fnm Þ2

where x is the liquid-phase mole fraction and g represents the activity coefficients. The subscripts “m” and “n” denote the refrigerant and the IL, respectively. F is an adjustable binary interaction parameter and s are defined in Equations (4) and (5).

Fmn hexpðusmn Þ

smn hs0mn þ

s1mn T

and and

Fnm hexpðusnm Þ

snm hs0nm þ

(4)

s1nm

(5)

T

where T is absolute temperature, respectively, and u ¼ 0:2. The parameters, s0mn , s1mn , s0nm , and s1nm , have been determined based on literature vaporeliquid equilibrium (VLE) data, as shown in Table 3. It is assumed that upon mixing, the refrigerant/IL mixtures are only physically interacting. To the best knowledge of the authors, no chemical reactions between the substances have been reported. The predicted mole fractions using the NRTL model are compared with the measured mole fraction data in references [38e42] in Fig. 2, which shows good agreement between the measured and predicted data. The correlations based on group contribution methods were used to evaluate the viscosity [43], specific heat [44] and density [45] of the ILs. The REFPROP 6.0 software [46], developed by the National Institute of Standards and Technology, has been used to calculate the thermodynamic and transport properties of the refrigerants. It implements three models for the thermodynamic properties of pure fluids: (i) equations of state explicit in the Helmholtz energy, (ii) the modified BenedicteWebbeRubin equation of state, and (iii) an extended corresponding states (ECS) model [46]. The specific enthalpy, hIL, and specific entropy, sIL, of ILs were calculated using Equations (6) and (7) [47].

ZT hIL ¼

ZP cp;IL dT þ

T0

ZT sIL ¼

3. System analysis

1007

vIL dP þ h0

(6)

P0

cp;IL dT þ s0 T

(7)

T0

The correlations of the refrigerant/IL mixture properties developed based on non-random two-liquid (NRTL) activity coefficient Table 2 Main components and the processes of a waste-heat driven absorption refrigeration system for electronic cooling applications. Components

States

Process

Evaporator

1/2

Absorber

2, 10 / 5

Pump Solution HX Desorber

5/6 6 / 7, 8 / 9 7 / 8, 3

Condenser Expansion device

3/4 4 / 1, 9 / 10

Heat absorption from an electronic device Refrigerant absorption/ condensation into IL Isentropic pressurization Regenerative pre-heating Refrigerant desorption / vaporization from IL Heat rejection to ambient Isenthalpic expansion

Table 3 Adjustable parameters in Equation (4). Working fluid pair (1)/(2)

s0mn [-]

s1mn [K]

s0nm [-]

s1nm [K]

R134a/[emim][Tf2N] [39] R134a/[bmim][PF6] [39] R134a/[hmim][Tf2N] [39] R134a/[hmim][BF4] [39] R134a/[hmim][PF6] [39] R32/[bmim][BF4] [38] R32/[bmim][PF6] [38] R125/[bmim][PF6] [38] R143a/[bmim][PF6] [38] R152a/[bmim][PF6] [38] R114/[emim][Tf2N] [40] R124/[emim][Tf2N] [40] Water/[emim][Tf2N] [41] Water/[emim][BF4] [42]

6.6710 1.2510 13.186 7.5975 11.718 0.93194 6.1356 2.4582 5.2848 6.5351 2.0631 0.11312 1.7388 17.253

716.04 411.45 2904.5 1176.7 2397.7 553.36 1194.6 48.172 18.277 871.79 1549.9 1210.1 1074.7 6445.4

0.8502 0.57596 5.3330 0.26344 3.5270 0.36807 1.9069 1.6394 0.11097 0.02731 1.0567 1.4565 4.9829 12.650

262.85 406.43 1128.9 275.97 688.02 585.91 122.26 563.00 68.435 409.50 390.84 1040.3 1369.1 4426.8

1008

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

Mole fraction predicted

1

_r ¼ m

0.8

-5% 0.4 0.2 0 0

0.2 0.4 0.6 0.8 Mole fraction measured

1

Fig. 2. Comparison between the measured [38e42] and the predicted mole fractions using the NRTL model.


 1  xm w _ _ 
ms ¼ mr m xs  xm w

(16)

_ sm _r _w ¼ m m

(17)

The subscripts “s” and “w” represent the strong- and weakrefrigerant solutions, respectively. Also, by assuming thermal equilibrium at the desorber outlet, T3 ¼ T8 , the heat transfer rate at the condenser can be calculated from Equation (18).

_ r ðh3  h4 Þ Qc ¼ m where cp,IL and vIL are the specific heat and specific volume of the IL, and P is the pressure. The IIR reference state was adopted; h0 ¼ 200 kJ=kg and s0 ¼ 1 kJ=kgK for the saturated liquid at T0 ¼ 0o C ðP0 ¼ Psat ðT0 ÞÞ. Then, the mixture enthalpy, H, and entropy, S, were calculated from Equation (8).

H ¼ H

id

þH

E

and

id

S ¼ S þS

E

where the ideal solution properties of enthalpy, Sid , can be obtained from Equation (9).

Hid ¼ xm Hm þ xn Hn

and

(8) H id ,and

entropy,

Sid

¼ ðxm Sm þ xn Sn Þ  Rðxm lnxm þ xn lnxn Þ

(15)

The maximum desorber outlet temperature, T8 , and the absorber outlet temperature, T5 , are adjustable parameters which m determine the refrigerant mass fractions of the strong ðxm s ¼ x5 Þ m m and weak solutions ðxw ¼ x8 Þ. The solution mass flow rates can be determined from Equations (16) and (17).

5%

0.6

ðh2  h1 Þ Qe

(9)

(18)

The pumping process is assumed to be isentropic, yielding Equation (19).

s6 ¼ s5

(19)

The energy balances at the solution heat exchanger, assuming no heat losses to the environment, is expressed by Equation (20).

_ s ðh7  h6 Þ ¼ m _ w ðh8  h9 Þ Qshx ¼ m

(20)

where Qshx is the heat transfer from the high temperature weak solution to the low temperature strong solution in the heat exchanger. The energy balances at the desorber and absorber is given by Equations (21) and (22).

The excess enthalpy, H E , and excess entropy, SE , can be calculated from Equations (10) and (11) [38].

_ w h8 þ m _ r h3  m _ s h7 Qd ¼ m

(21)

      vlngm vlngn HE ¼ RT 2 xm þ xn vT vT P;x P;x

(10)

_ w h10 þ m _ r h2  m _ s h5 Qa ¼ m

(22)

TSE ¼ HE  GE

(11)

where R is the universal gas constant and The excess Gibbs energy, GE, can be calculated from Equation (12).

GE ¼ xm lngm þ xn lngn RT

(12)

Since the feasibility and the compatibility of ILs as an absorbent in an absorption system, in combination with refrigerants are the principal foci in this study, the performance of the system with respect to the refrigerant/IL mixtures is thermodynamically evaluated as a function of operating conditions. Given the degree of subcooling at the condenser outlet, degree of superheat at the evaporator outlet, and the saturation temperatures for the condenser and evaporator, the condenser and evaporator outlet states (states 4 and 2) can be determined. Isenthalpic throttling through both the refrigerant and solution expansion devices is assumed resulting in Equations (13) and (14).

h1 ¼ h4

(13)

h10 ¼ h9

(14)

The subscript numbers denote the states indicated in Fig. 1. The energy balance at the evaporator yields the refrigerant mass flow rate, Equation (15).

where Qa is the heat removed from the refrigerant/IL mixture at the absorber. The viscosity of ILs are often relatively high; for example, for [bmim][PF6] at 294 K and atmospheric pressure it is 376 mPa s [48]. Thus, the IL flow in micro-channels used in the absorber and desorber may create large pressure drops, which can affect the system performance. The average pressure drops through the microfluidic channel heat exchangers were evaluated using a twophase pressure drop equation, Equation (23).

" " # #   v 2 dP 2fl G2m ð1  xv Þ 2 ð1  xv Þ2 2 d ðx Þ fl þ Gm ¼  þ dz dh rl dz εrv ð1  εÞrl

(23)

where dh is the hydraulic diameter of channel, and f, G, xv, r, and ε are the liquid-phase fanning friction factor, mass flux, vapor quality, density, and void fraction, respectively. z is the axial direction coordinate along the channel length. Subscripts “l” and “v” represent liquid and vapor phase, respectively. In the two-phase multiplier correlation of Lockhart and Martinelli [49], fl is incorporated with the C value proposed by Lee and Mudawar [50], Equations (24)e(26).

C X

f2l ¼ 1 þ þ

1 X2

 C ¼ 2:16 Re0:047 We0:6 laminar liquid  laminar vapor lo lo

(24)

(25)

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

laminar liquid  turbulent vapor

(26)

where Relo and Welo are liquid-only Reynolds and Weber numbers, respectively. The Martinelli parameter, X, and the single phase empirical correlation of fanning friction factor for laminar flow in a rectangular channel by Shah and London [51] are expressed by Equations (27) and (28).

 X ¼

ml mv

0:5 

1  xv xv

0:5  0:5

rv rl

(27)

a

0.5

T8,min ∼ 61.5oC

R32

0.3 R152a

0.2 0.1

 2 3 4 f Re ¼ 24 1  1:3553b þ 1:9467b  1:7012b þ 0:9564b 5 ð28Þ  0:2537b

0 60

70

80

90

100

Desorber outlet temperature [oC]

b

0.5

T8,min ∼ 61.5 C o

0.4 COP(Qe/(Qd+Wp))

where m is the viscosity and b is the aspect ratio of the channel. Also, the void fraction model of Zivi [52] is adopted in this study. Microchannel structures for the absorber and the desorber are adopted as a baseline, due to their ability to yield high heat and mass transfer rates and to minimize transport limitations on the performance; yet, the extent of potentially negative effect of the high viscosity of ILs on the pressure drop in microchannel heat/ mass exchangers and, in turn, the system performance is critically assessed. The dimensions (length  width) of the evaporator and condenser are 2  2 cm, and 3  3 cm, respectively. The dimensions of the absorber and the desorber are 8  8 cm. The microfluidic channel cross-sectional area for the heat exchangers is 1  1 mm.

1009

0.4 COP(Qe/(Qd+Wp))

0:23 C ¼ 1:45 Re0:25 lo Welo



0.3 0.2 0.1

water

4. Results & discussion

0

4.1. Effect of the desorber outlet temperature

60

xw w

Pc

gw Psat ðT8 Þ

(29)

where Pc is the vapor pressure at the condenser and Psat ðT8 Þ is the saturation vapor pressure at T8 . Subscripts “c” and “sat” represent the condenser and saturation, respectively. As the temperature increases, the refrigerant mole fraction decreases. The flow rate of the solution will then vary correspondingly. The circulation ratio, a, is the amount of solution flow rate of the refrigerant/IL mixture needed to achieve compression for a given mass flow rate of the refrigerant, as given by Equation (30).


 _ 1  xm m ah s ¼
m wm  _r m x s  xw

(30)

The circulation ratio increases with mass fraction in the weak solution, at fixed refrigerant mass fraction in the strong solution. Fig. 4 shows the circulation ratio variation with respect to the desorber outlet temperature. The desorber outlet temperature rise reduces the circulation ratio due to a decrease in the mass fraction of the weak solution. Equation (20) can be rewritten to yield Equation (31).

80

90

100 o

Desorber outlet temperature [ C]

c COP (Qe/(Qd+Wp))

Fig. 3 shows the COP variation of the absorption system with respect to desorber outlet temperature for several working fluid pairs. Fig. 3 shows a relatively sharp rise in COP up to w80
C followed by a more constant value as the desorber outlet temperature rises beyond 80
C. The variation of the desorber outlet temperature affects the concentration of refrigerant in the weak solution, which can be approximated by Equation (29) [47].

70

1 0.8 0.6

T8,min ∼ 61.5 oC

0.4

R32/[bmim][BF4]

0.2 R134a/[hmim][BF4]

0

60

70

80

90

100 o

Desorber outlet temperature [ C] Fig. 3. COPs of the absorption system with respect to the desorber outlet temperature using different working fluid pairs: (a) refrigerant/[bmim][PF6] and R134a/[hmim] [PF6]; (b) refrigerant/[emim][Tf2N] and R134a/[hmim][Tf2N]; (c) R134a/[hmim][BF4], R32/[bmim][BF4] and water/[emim][BF4].

Qd ¼ aðh8  h7 Þ  h8 þ h3 mr

(31a)

Qd ¼ h8 ða  1Þ þ h3  ah7 mr

(31b)

1010

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

rate of the circulation ratio ðva=vT8 Þ is diminished (Fig. 4). Also, the increase in enthalpies of the desorber outlet (h3 and h8 ) accompanied by the increase of the temperature require larger heat input to the desorber, Equation (31b). Consequently, at high desorber outlet temperatures, the effect of the reduced circulation ratio and the increased enthalpy on the heat input to the desorber conflicts and the effect on COP cancels each other out. This leads to a leveling off of the COPs in Fig. 3. The refrigerant mole fraction for the strong solution is approximated in Equation (32) [47].

1000

α (Circulation ratio)

a

100

R125

10 xs w

R152a R134a

70

80

90

100 o

Desorber outlet temperature [ C]

b 1000

Pc Psat ðT5 Þ Pe

(33)

 Pc Psat ðT5 Þ wT8;min Pe

(34)

Psat ðT8 Þ>  T8 >Tsat

α (Circulation ratio)

(32)

Assuming gs zgw , Equations (29) and (32) can be written to yield Equations (33) and (34) using the inequality, xs > xw .

1 60

Pe

gs Psat ðT5 Þ

From Equation (34), it can be seen that the lower bound on the desorber outlet temperature, or maximum operating temperature of the system, is a function of the operating conditions, i.e., vapor pressures at the evaporator and condenser as well as the absorber outlet temperature. Fig. 3 also shows that the COP curves generally converge to zero at T8 w 61.5
C (dashed line) for all refrigerants. Thus, the COP curves can be shifted to lower temperatures by adjusting the operating conditions so as to have optimum performance at a lower temperature, ca. 60
C. Then, the utilization of waste-heat can be optimized.

100

10

1 60

70

80

90

100 o

4.2. Effect of refrigerant/IL compatibility

Desorber outlet temperature [ C]

α (Circulation ratio)

c

The R32/[bmim][PF6] pair showed the highest COP w0.42, of all the refrigerant pairs in Fig. 3(a). The COP values for R134a and R152a with [bmim][PF6] were moderate, while for R125 and R143a were lower. Note that the COP trends for the HFC refrigerants in Fig. 3(a) approximately correlate with the refrigerant mass fraction in the strong solution, xm s , listed in Table 4.Equation (35) shows the relation between the mass fraction and mole fraction.

1000

100

xm ¼

10

xMr xMr þ ð1  xÞMIL

(35)

where Mr and MIL are the molecular weights of refrigerant and IL, respectively. The effect of the strong solution solubility on the

Table 4 Refrigerant mass fractions in the strong solution.

1 60

70

80

90

100 o

Desorber outlet temperature [ C] Fig. 4. Circulation ratios of the absorption system with respect to the desorber outlet temperature using different working fluid pairs: (a) refrigerant/[bmim][PF6] and R134a/[hmim][PF6]; (b) refrigerant/[emim][Tf2N] and R134a/[hmim][Tf2N]; (c) R134a/ [hmim][BF4], R32/[bmim][BF4] and water/[emim][BF4].

Equation (31a) shows that a smaller circulation ratio results in less heat input to the desorber for the same amount of cooling at the evaporator, which would increase COP. Thus, a lower circulation ratio resulting from the temperature increase in Fig. 4 increases COP at low desorber outlet temperatures (<80
C), as shown in Fig. 3. However, as the temperature further increases, the change

Working fluid pair (1)/(2)

xm s

R143a/[bmim][PF6] R125/[bmim][PF6] R134a/[hmim][PF6] R152a/[bmim][PF6] R134a/[bmim][PF6] R32/[bmim][PF6] R114/[emim][Tf2N] Water/[emim][Tf2N] R134a/[hmim][Tf2N] R134a/[emim][Tf2N] R124/[emim][Tf2N] R134a/[hmim][BF4] R32/[bmim][BF4] Water/[emim][BF4]

0.0787 0.2210 0.2563 0.2286 0.3239 0.4362 0.0599 0.0091 0.2910 0.2676 0.3972 0.2791 0.4487 0.1542

[e]

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

vlnx2 vlnT

 ¼ P

D sr Rðvlna2 =vlnx2 ÞP;T

(36)

where Dsr is the entropy of mixing (or relative partial molar entropy) of refrigerant in the solution, and a2 is the activity of refrigerant in the solution given by Equation (37).

a2 ¼ g2 x2

(37)

In the Henry’s law regime (i.e., when g2 is independent of x2 ), the derivative in the right-hand side dominator in Equation (36) becomes unity. Therefore, the equation can be reduced to the van’t Hoff form [57] as follows,

  vlnx2 Dsr ¼ R vlnT P

Dsr R

lnT þ D

Refrigerant mass fraction ratio

0.61 R134a

0.37

R125

0.22 67

78

88

99

111

o

Desorber outlet temperature [ C]

b

1 R134a /[hmim][Tf2N]

0.61

R134a

0.37

0.22 57

67

78

88

99

111

Desorber outlet temperature [oC] (38)

c 1

Integrating Equation (38) yields [58]:

ðlnxÞP ¼

1

57

(39)

where D is an integration constant. Since the relative partial molar entropy is reduced by increasing the alkyl chain length, Equation (39) supports the reduced dependence of the refrigerant mass fraction on the desorber outlet temperature. In summary, longer cation alkyl chain length cause a larger solubility but lower dependence of the solubility on temperature. These effects of the cation result in larger circulation ratio for R134a/[hmim][Tf2N], as shown in Fig. 4(b), because the refrigerant mass fraction difference m between the strong and weak solutions in Equation (30), ðxm s  xw Þ, is lowered due to the reduced dependence of the solubility on temperature with the longer alkyl chain length of [hmimþ]. Thus, even with larger refrigerant solubility in the strong solution, the performances of the system can be degraded due to the reduced dependence of the solubility on temperature. R143a/[bmim][PF6] and R134a/[hmim][PF6] are remarkably less sensitive to temperature (Fig. 5(a)) and show relatively low COPs (Fig. 3(a)). Water solubility in [emim][Tf2N] depends relatively more on the desorber outlet temperature (Fig. 5(b)), but refrigerant mass fraction in strong solution is extremely low so that very large circulation ratio and thereby low COP are found in Fig. 4(b) and Fig. 3(b), respectively. Ren and Scurto [39] investigated the effect of R134a solubility on anion type in the IL and found that the solubility scales with the

Refrigerant mass fraction ratio



a

Refrigerant mass fraction ratio

performance is attributed to the smaller circulation ratio brought about by a larger value of refrigerant mass fraction in the strong solution. Thus, it can be stated that high solubility of refrigerant (mass fraction solubility) in the IL solution improves the performance of the absorption system. However, the higher solubility of R134a in [hmim][Tf2N] results in a lower COP value relative to R134a/[emim][Tf2N] in Fig. 3(b). Ren and Scurto [39] compared the solubility (mole fraction solubility) of R134a in [emim][Tf2N] and R134a in [hmim][Tf2N]. They reported that a longer alkyl chain length on the cation (i.e. hmimþ vs. emimþ) increases the solubility of R134a. Shiflett and Yokozeki [38] and Aki et al. [53] measured the solubility of R32 and carbon dioxide in the ILs, respectively. They both also found that the solubility of the gas in the IL can be increased by increasing the alkyl chain length on the organic cation, which is in agreement with the results in Fig. 5(b). On the other hand, Kerle et al. [54] investigated the temperature dependence of the solubility of carbon dioxide in imidazolium-based ILs and showed that as the alkyl chain length increases, the solubility becomes less sensitive to temperature. The temperature coefficient of solubility is directly related to the entropy of mixing in Equation (36) [55e58].

1011

0.61

0.37

0.22

0.14 57

67

78

88

99

111 o

Desorber outlet temperature [ C] Fig. 5. Refrigerant mass fraction ratio in the weak solution relative to the strong m solution ðlnxm w =xs Þ of the absorption system as function of the desorber outlet temperature (lnT8 ), using different working fluid pairs: (a) refrigerant/[bmim][PF6] and R134a/[hmim][PF6]; (b) refrigerant/[emim][Tf2N] and R134a/[hmim][Tf2N]; (c) R134a/ [hmim][BF4], R32/[bmim][BF4] and water/[emim][BF4].

ionic radius of the anion or charge density. A lower charge density allows the polar hydrofluorocarbons to more readily dissolve. Fig. 5(aec) shows that the solubility of R134a in [hmim][PF6] and [hmim][BF4] are similar while R134a is more soluble in [hmim] [Tf2N] (Table 4). Therefore, the high solubility of R134a in [hmim]

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

[Tf2N] and the high temperature dependence of the R134a solubility in [hmim][BF4] yield higher COP values than that of R134a/ [hmim][PF6]. The differences in both the solubility (Fig. 5(a) and (c)) and COP values (Fig. 3(a) and (c)) for R32 in [bmim][PF6] and [bmim][BF4], were not significant. The COP values for the absorption system using the water/ [emim][BF4] pair are notably higher than those with the other working fluid pairs (Fig. 3(c)). The maximum COP was 0.91 at Td ¼ 70
C. The imidazolium-[BF 4 ] ILs have dramatically higher solubility for water (Fig. 4(c)), while ILs with [Tf2N] show a decrease in water solubility [26] as shown in Fig. 4(b). The low COP values of water/[emim][Tf2N] in Fig. 3(a) are, therefore, attributed to the low solubility of water in [emim][Tf2N]. However, the refrigerant mass fraction of water in [emim][BF4]is less than that of R32 (Fig. 5(c)). Moreover, the circulation ratio of water is the highest among the refrigerants for [emim][BF4] in Fig. 4(c), by which the lower COP is expected than that of R32 in [emim][BF4]. This suggests that there should be another factor affecting the performance of the absorption system, other than solubility and circulation ratio. The latent heat of vaporization for water is an order of magnitude higher than that of the other refrigerants as listed in Table 1. Then, the required mass flow rate of water to produce the target 100 W cooling capacity at the evaporator is significantly smaller (Table 5). It is evident from Equation (31) that the low water flow rate contributes to reducing the heat input required at the desorber, Qd . Consequently, this leads to high COP values for the water/[emim][BF4] pair shown in Fig. 3(a). Similarly, the refrigerants R152a and R32 also have large latent heats and small refrigerant mass flow rates. This improves the system performance when using these working fluid pairs. 4.3. Effect of IL viscosity Higher viscosity ILs cause an increased pressure drop in the compression loop, which would result in larger pumping power or larger pipes and system volume. Fig. 6 shows the viscosity of pure ILs vs. temperature evaluated using the group contribution method [43]. The viscosity increases with cation mass: emimþ < bmimþ < hmimþ. The viscosity is more dependent on the  anion with the following order: Tf2N < BF 4 < PF6 . The viscosity of [emim][Tf2N] is only 31.3 mPa s at 294 K, which is 10 times smaller than that of [hmim][PF6]. Fig. 7 shows the total pumping power of the absorption system for various refrigerants/IL pairs. It is very encouraging that for most of the working fluid pairs, except R125/[bmim][PF6], R143a/[bmim] [PF6] and R114/[emim][Tf2N], the pumping power was less than 10 W for the 100 W cooling capacity system. The total pumping power can be calculated using Equations (40) and (41).

 _ s Pc  Pe þ DPfriction Wp ¼ m

(40)

Table 5 Refrigerant mass flow rates in the absorption system. Refrigerants

Refrigerant flow rate [g/s]

R114 R124 R125 R134a R143a R152a R32 Water

0.897 0.785 1.155 0.651 0.761 0.402 0.420 0.042

10 [bmim][BF4]

Viscosity [Pa s]

1012

1

[hmim][BF4] [bmim][PF6] [hmim][PF6]

0.1

0.01

[emim][Tf2N] [emim][BF4 ]

0.001 250

[hmim][Tf2 N]

300 350 Temperature [K]

400

Fig. 6. Viscosities of ILs evaluated using group contribution method [43].

 Wp ¼ a Pc  Pe þ DPfriction _ mr

(41)

where DPfriction is the pressure drop due to friction. The circulation ratio plays an important role in the pumping power, as well. Since a lower solution flow rate is accompanied by a reduction in the circulation ratio at higher desorber temperatures, as shown in Equation (40), lower pumping power values result from a higher desorber temperature. Also, COPwaste (ratio of the evaporator cooling capacity to the pumping power) improves as the pumping power decreases at higher desorber temperatures, as shown in Fig. 8. Due to the small flow rate of _ r , the pumping power becomes negligible, resulting in water, m very high COPwaste values for the water/[emim][Tf2N] and water/ [emim][BF4] pairs. However, it should be noted that even though the COPwaste values are large for water/[emim][Tf2N], the low value of regular COP requires large desorber and absorber heat exchangers. Fig. 9 shows the ratio of pumping power (due to friction) to the total pumping power. For most working fluid pairs, less than 10% of the total pumping power is attributed to friction. Most of the pumping power is used to pressurize the refrigerant/IL pair ðPc  Pe Þ, which is the reason that R32 shows the smallest frictional contribution while water has the largest. Since the pressure difference between the condenser and evaporator with water is only 9.2 kPa, which is negligibly small and less than 1% of the pressure difference with R32, more than 90% of the pumping power for the water/[emim][Tf2N] and water/ [emim][BF4] was used to overcome the frictional losses. For the R143a/[bmim][PF6] and R114/[emim][Tf2N] pairs, the high solu_ s , resulted from high circulation ratios and, in tion flow rate, m turn, caused large values of pumping power due to frictional losses. The relatively small contribution of friction to the total pumping power was mainly because of the low viscosity due to dilution of the refrigerants. However, it also should be noted that the heat exchanger microchannel cross-sectional area was slightly larger than that of conventional microfluidic channels. As shown in Equation (28), the laminar flow friction factor is inversely proportional to the Reynolds number, fl wRe1 wm=Gdh . Under fixed mass flow rate, mass flux is inversely proportional to the channel cross2 sectional area, GwA1 cs w4=pdh , where Acs is the channel cross-

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

Total pumping power , Wp[W]

a

1013

a

1000

100

10

R125 R32

R134a

1 60

70

80

90

100 o

Desorber outlet temperature [ C]

Total pumping power, Wp[W]

b

b

1000

100

R114

10

1

0.1 60

70

80

90

100

Desorber outlet temperature [oC]

Total pumping power,Wp[W]

c

c

100

10

1

0.1

0.01 60

70

80

90

100 o

Desorber outlet temperature [ C] Fig. 7. Total pumping power consumptions of the absorption system with respect to the desorber outlet temperature using different working fluid pairs: (a) refrigerant/ [bmim][PF6] and R134a/[hmim][PF6]; (b) refrigerant/[emim][Tf2N] and R134a/[hmim] [Tf2N]; (c) R134a/[hmim][BF4], R32/[bmim][BF4] and water/[emim][BF4].

Fig. 8. COPwaste of the absorption system with respect to the desorber outlet temperature using different working fluid pairs: (a) refrigerant/[bmim][PF6] and R134a/[hmim][PF6]; (b) refrigerant/[emim][Tf2N] and R134a/[hmim][Tf2N]; (c) R134a/ [hmim][BF4], R32/[bmim][BF4] and water/[emim][BF4].

 sectional area. Assuming that the two-phase multiplier, f2l ¼ ðdP=dzÞTP =ðdP=dzÞl , is a weak function of diameter, Equation (23) shows that the frictional pressure drop, which is the first bracket of the right-hand side of the equation, can be given by Equation (42).

dP dz



 ¼ friction

 2fl G2 ð1  xv Þ G f2l w 2 wd4 h dh rl dh

(42)

Thus, small adjustments in the microfluidic channel dimensions can have a large effect on the pumping power for systems with high viscosity.

1014

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

a

COP of the system, as shown in Fig. 10. As the absorber outlet temperature rises, the solubility of the strong solution decreases, as determined by Equation (32). Thus, the difference in mole fraction between the strong and weak solutions is diminished, which leads to a larger circulation ratio and lower COP. The increase in the desorber inlet enthalpy, h7 , driven by the absorber outlet temperature increase, also contributes to the enhancement of COP by reducing the required heat input to the desorber, Qd . However, the

a

b

b

c

c

Fig. 9. Pumping power consumptions of the absorption system caused by friction with respect to the desorber outlet temperature using different working fluid pairs: (a) refrigerant/[bmim][PF6] and R134a/[hmim][PF6]; (b) refrigerant/[emim][Tf2N] and R134a/[hmim][Tf2N]; (c) R134a/[hmim][BF4], R32/[bmim][BF4] and water/[emim][BF4].

4.4. Effect of the absorber outlet temperature The absorber outlet temperature is controlled by the ambient cooling temperature. A lower absorber temperature enhances the

Fig. 10. COPs of the absorption system with respect to the absorber outlet temperature using different working fluid pairs: (a) refrigerant/[bmim][PF6] and R134a/[hmim] [PF6]; (b) refrigerant/[emim][Tf2N] and R134a/[hmim][Tf2N]; (c) R134a/[hmim][BF4], R32/[bmim][BF4] and water/[emim][BF4].

Y.J. Kim et al. / Energy 44 (2012) 1005e1016

effect of the circulation ratio change on COP dominates, so that the monotonic decrease in COP, with respect to the absorber temperature increase, is observed in Fig. 10. 5. Conclusion The feasibility of several refrigerant/IL mixtures as working fluid pairs for the absorption based refrigeration system was numerically evaluated. The refrigerant/IL compatibility indicated by the circulation ratio was one of the most important factors in determining the system performance with the new working fluid pairs. The solubility of the refrigerants in the ILs and their temperature dependence affected the circulation ratio. R32 showed the lowest circulation ratio with relatively high COP values for the ILs  having [PF 6 ] and [BF4 ] anions. Although R124 showed the best compatibility and performance with [Tf2N], due to its negative environmental impact, R124 may not be the optimum choice of refrigerant. Cations with shorter alkyl chains are preferred ([emimþ] > [bmimþ] > [hmimþ]) due to more sensitive dependence of the solubility on temperature. Overall, water/[emim][BF4] showed the highest COP value, ca. 0.91, where the better compatibility of water with [emim][BF4] and the superior properties of water as a heat transfer fluid, such as large latent heat of evaporation, followed by extremely small refrigerant (water) flow rate resulted in its high performance. Using water with [emim][Tf2N] has not been found promising due to the extremely low solubility of IL in water (Table 4). Lafrate et al. [59], however, showed that diolfunctionalized ILs with [Tf2N] can be completely water-miscible, which could potentially deliver an improved performance of the absorption system. However, the evaporator and condenser temperature and pressure values for water (as the refrigerant) are not of interest for most electronic cooling applications. The effect of the higher viscosity of the ILs on the pressure drop and pumping power were mitigated by slightly increasing the microfluidic channel cross-sectional area. The absorber outlet temperature significantly affected the system performance. Lowering the absorber outlet temperature is desirable due to the increased solubility of the refrigerant in the IL. The system level numerical investigation showed that refrigerant/IL pairs are promising materials for absorption refrigeration utilizing low-grade waste heat, as found in electronic systems. For practical implementation of ILs in the absorption systems the complimentary studies concerning the detailed effects of heat and mass transfer characteristics and geometries of heat/mass exchangers are needed. Also, optimal microfluidic absorber and desorber configurations need to be developed to fully realize the promising potential of an absorption refrigeration system based on the IL chemical compressor. Acknowledgments The authors acknowledge the support of the Interconnect Focus Center, one of the five research centers funded under the Focus Center Research Program, a Semiconductor Research Corporation and DARPA program. Nomenclature C COP COPwaste cp dh F f

coefficients in Lockhart and Martinelli correlation coefficient of performance coefficient of performance with waste-heat specific heat [J/kg K] hydraulic diameter [m] adjustable binary interaction parameter fanning friction factor

G Gm H h M _ m P Q R Relo S s T v Welo Wp X x xm xm xn xv z

1015

Gibbs energy [J] mass flux [kg/m2s] enthalpy [J] specific enthalpy [J/kg] molecular weight [kg/mol] mass flow rate [kg/s] pressure [Pa] heat transfer [W] gas constant, 8.314 J/mol K liquid-only Reynolds number entropy [J/K] specific entropy [J/kg K] temperature [K] specific volume [m3/kg] liquid-only Weber number pumping work [W] Martinelli parameter liquid phase refrigerant mole fraction liquid phase refrigerant mole fraction liquid phase refrigerant mass fraction liquid phase ionic-liquid mole fraction vapor quality coordinate [m]

Greek symbols b aspect ratio ε void fraction 4 two-phase multiplier g activity coefficient m viscosity [Pa s] r density [kg/m3] s adjustable parameters of Equation (5) Subscripts 0 reference state 1, ., 10 state numbers indicated in Fig. 1 a absorber c condenser d desorber e evaporator IL ionic-liquid l liquid r refrigerant s strong-refrigerant solution sat saturation shx solution heat exchanger v vapor w weak-refrigerant solution Superscripts E excess property id ideal solution References [1] International technology roadmap for semiconductors, assembly and packaging, 2005 Ed. [2] Pal A, Joshi YK, Beitelmal MH, Patel CD, Wenger TM. Design and performance evaluation of a compact thermosyphon. IEEE Trans Comp Pack Tech 2002; 25(4):601e7. [3] Maydanik YF, Vershinin SV, Korukov MA, Ochterbeck JM. Miniature loop heat pipes-a promising means for electronics cooling. IEEE Trans Comp Pack Tech 2005;28(2):290e6. [4] Jiang L, Mikkelsen J, Koo JM, Huber D, Yao S, Zhang L, et al. Closed-loop electroosmotic microchannel cooling system for VLSI circuits. IEEE Trans Comp Pack Tech 2002;25(3):347e55. [5] Wei Y, Joshi YK. Stacked microchannel heat sinks for liquid cooling of microelectronic components. J Electron Pack 2004;126:60e6.

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[33] [34] [35] [36]

[37]

[38] [39]

[40] [41]

[42]

[43]

[44] [45]

[46]

[47] [48]

[49] [50]

[51]

[52]

[53]

[54]

[55] [56] [57]

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