九州大学大学院総合理工学報告第32巻 第3号 12-18頁 平成22年11月

Engineering Sciences Reports, Kyushu University
Vol. 32, No. 3, pp. 12-18, NOV, 2010

Performance Investigation of a Novel CO2Compression-Adsorption Based Hybrid Cooling Cycle
Skander JRIBI *1,† Shigeru KOYAMA*1 Bidyut Baran SAHA*2
E-mail of corresponding author: s.jribi@cm.phase.kyushu-u.ac.jp†(Received Nov. 01, 2010)
In this paper, the performance of a CO2 based hybrid compression-adsorption cooling cycle ispresented based on a transient mathematical model. In the proposed hybrid cycle two stage compression have been performed where CO2 passes through the mechanical compressor at thelow-pressure side. The compressed CO2 from the mechanical compressor is then adsorbed in thedesignated adsorbers of the adsorption cycle. The chiller response as well as the chiller performancein terms of cooling capacity and coefficient of performance (COP) are investigated for different desorption pressures and gas cooler outlet temperatures. This novel hybrid cooling system attains
similar cooling power and two to three folds COP values than those of the conventional CO2 based
vapor compression cycle with the same operating conditions. Therefore, the proposed system seems
to be promising for automotive air-conditioning application where waste heat from engine is esteemed.
Key words: COP, Cooling capacity, Activated carbon, CO2, Hybrid, Chiller

1.

Introduction

Nowadays, R134a is the most used refrigerant in automotive and domestic air conditioning systems. However, the global warming
potential (GWP) of R134a is as high as 1300.
So it is essential to replace R134a with an environmentally benign refrigerant. CO2 has theadvantages of being natural, non-flammableand non-toxic refrigerant. Moreover, CO2 has azero ODP (Ozone Depletion Potential) and anegligible GWP equal to one. Recent studies
show that CO2 have similar performance tothat of the R134a in automotive and domestic
air conditioning applications.1,2) However, CO2vapor compression cooling systems suffer fromhigh working pressures which led to 10 to 30%
higher manufacturing costs in comparing tothose of the R134a based systems.3) So, it isinevitable to improve the performance of CO2
cooling systems.
In this study, a novel hybrid compression-adsorption cooling system is proposed
where the compression process is realized in
two steps: (i) mechanical compression at thelow pressure side and (ii) thermal compression*1*2

Department of Energy and Environmental
Engineering
Department of Mechanical Engineering, Faculty ofEngineering

by adsorption/desorption processes at the highpressure side. This innovative hybrid system
uses highly porous activated carbon of type
Maxsorb III and CO2 as the adsorbent-refrigerant pair. The paper also presents
the transient mathematical model and theperformance results of the hybrid cycle which
are determined by cycle simulation.

2.

Description of the Hybrid Cooling
Cycle

Fig.1 shows the schematic diagram of thehybrid compression-adsorption cooling system
which consists of: a gas cooler, a liquid
reservoir, an expansion valve, an evaporator, amechanicalcompressor
and4
adsorber/desorber beds filled with Maxsorb III.
At first, the refrigerant, CO2 is evaporated in
the evaporator. Then, the evaporated
refrigerant is compressed with the mechanicalcompressor and refrigerant pressure increases
from Pevap to Pcomp. After that, CO2 is adsorbed
in the beds assigned in adsorption mode
known as adsorption process. Adsorption heatis generated during the adsorption process,which is removed by the cooling water. In thenext step, the adsorber bed is disconnectedfrom the mechanical compressor and it isheated by the hot water and this is known as

13

Performance Investigation of a Novel CO2 Compression-Adsorption Based Hybrid Cooling Cycle

pre-heating process. The pressure in the bed
during the pre-heating process increases until
it reaches the gas cooler pressure (Pgc).
Concurrently, the valve between the bed andthe gas cooler is opened and CO2 starts todesorb from the desorbing bed. Desorption
heat is absorbed during the desorption process,which is compensated by the heating water.
The refrigerant desorbed from activatedcarbon is then cooled down in the gas cooler
with the cooling water. Finally, CO2 isexpanded with the expansion valve.
As the process of adsorption/desorption isintermittent, four beds are employed to have acontinuous cycle. The four beds works indifferent processes as shown in table 1. A, C, D
and H denotes the processes of adsorption,
pre-cooling, desorption and pre-heating,respectively. For example, in mode ADDA, beds
1,4 are in adsorption phase and beds 2,3 are indesorption phase. In mode ACDH, beds 1, 2, 3
and 4 are in adsorption, pre-cooling, desorption
and pre-heating phases, respectively.

3.

3.1 Adsorption isotherms
The equilibrium amount adsorbed of CO2
onto Maxsorb III had been investigated experimentally4) and Tόth equation (eq. 1) is used
to correlate the adsorption data.

⎛ Q ⎞b0 P exp⎜⎟⎝ RT ⎠⎛ ⎛⎞⎜1 + ⎜ b P exp⎛⎜ Q ⎞⎟ ⎞⎟ ⎟
0
⎜
⎟
⎜ ⎝
⎝ RT ⎠ ⎠ ⎟⎠
⎝
t

3.2

Schematic diagram of vapor compression-adsorption hybrid cooling system. Beds 1,4
are in desorption phase and beds 2,3 are in adsorption phase.

⎛ ∂ ln P ⎞Qst = RT 2 ⎜⎟⎝ ∂T ⎠ C(2)

The heat of adsorption derived from Tόthisotherm equation is constant and equals theparameter Q of eq. 1.

Mathematical Model

C = C0

Fig.1

(1)

1
t

Heat of adsorption

The heat of adsorption is computed from theClausius-Clapeyron equation given by:

3.3

Adsorption kinetics

The adsorption rate is described by the linear driving force theory given by:dc= k s av (C − c)dt(3)where C is the equilibrium uptake, c stands for
instantaneous uptake and ksav is the overall
mass transfer coefficient estimated experimentally for the adsorption of CO2 onto activated carbon beads.5)⎛ 1915⎞k s av = exp⎜ −+ 1.776 ⎟
⎝ T
⎠
3.4

(4)

Adsorbing bed energy balance

The rate of change of adsorbing bed internal
Table 1 Modes of adsorption/desorption processes
Mode

ADDA

Bed1
Bed2

Des.

Bed3
Bed4

Ads.

ACDH

AADD

HACD

DAAD

DHAC

Adsorption

Pre-Heating

Desorption

Pre-Cooling

Adsorption

Pre-Heating

Desorption

Pre-Cooling

Adsorption

Pre-Heating

Desorption

Pre-Cooling

DDAA

CDHA
Pre-Cooling

Desorption
Pre-Heating
Adsorption

平成22年度

九州大学大学院総合理工学報告 第32巻 第3号

energy is equal to the heat released by adsorption process Qads minus the heat removed by
the cooling water Qcool and this is written as:

dTbed,ads

dcads
⋅
(5)
dt
dt
(Qst +hbed,ads −hcomp,out)−(m&cp )cool(Tcool,out −Tcool,in)(Mcp )bed= MMX3where the cooling water outlet temperature ismodeled by the log mean temperature difference (LMTD) method and expressed by:⎛ − UAbed ⎞⎟Tcool,out = Tbed,ads + (Tcool,in − Tbed,ads ) exp⎜
⎜ (m& c p ) ⎟cool ⎠⎝(6)
3.5

Desorbing bed energy balance

The rate of change of desorbing bed internalenergy is equal to the heat absorbed by desorption process Qdes minus the heat added by
the heating water Qhot and this is expressed as:

dTbed ,desdc
= M MX 3 des Qst
dtdt&− (mc p ) hot (Thot ,in − Thot ,out )( Mc p ) bed(7)where the hot water outlet temperature is determined by:⎛ − UAbed ⎞⎟Thot ,out = Tbed ,des + (Thot ,in − Tbed ,des ) exp⎜
⎜ (m& c p ) ⎟
hot ⎠
⎝
(8)
3.6 Evaporator energy balance
The change of evaporator’s internal energy
is described by eq. 9. The first term in the righthand side (R.H.S.) of eq. 9 represents the latent heat of evaporation of the liquid refrigerant and the second term in the R.H.S. of eq. 9
represents the cooling capacity Qchill.

dTevapdcads(hevap,out − hevap,in )dt
dt− (m& c p ) chill (Tchill,out − Tchill ,in )( Mc p ) evap

= M MX 3(9)
The chilled water outlet temperature isgiven by:⎛ − UAevap ⎞⎟Tchill ,out = Tevap + (Tchill ,out − Tevap )exp⎜⎜ (m& c p ) ⎟chill ⎠
⎝

(10)

14

The evaporator’s overall heat transfer coefficient (UAevap) is estimated for an average flow
boiling heat transfer of CO2 in round-tube.6)
3.7 Gas cooler energy balance
The change of gas cooler internal energy isequal to the heat removed from the desorbed
refrigerant minus the heat released to thecooling water Qcw and this is given by:

dTgc

dcdes
((hbed,des − hgc ))dt
dt& c p )cw (Tcw,out − Tcw,in )− (m(Mcp ) gc= M MX3(11)

where the cooling water outlet temperaturefrom gas cooler is given by:⎛ − UAgc ⎞⎟Tcw,out = Tgc + (Tcw,in − Tgc )exp⎜⎜ (m& c p ) ⎟
cw ⎠
⎝

(12)

The gas cooler’s overall heat transfer coefficient (UAgc) is estimated for an average forced
convection cooling heat transfer of CO2 in microchannel tube.6)
3.8

Compressor energy balance

The mechanical work consumed by thecompressor is expressed by:

dc
W&Comp = M MX 3 ads (hcomp ,out − hcomp ,in )
dt

(13)

An isentropic efficiency (ηis) of the compressor is assumed to be 0.8.
3.9

Mass balance

The mass balance implies that the adsorbed
amount in the activated carbon is equal to themass desorbed:

mads = mdes

(14)

where the masses adsorbed and desorbed
are given by the following equations, respectively.

dmads
dc
= M MX 3 ads
dt
dt

(15)

dmdes
dc
= M MX 3 des
dt
dt

(16)

3.10

Energy balance

The energy balance (eq. 17) implies that theheats adsorbed by evaporator and desorption
beds and the mechanical power consumed bycompressor are equal to the heats rejected by15

Performance Investigation of a Novel CO2 Compression-Adsorption Based Hybrid Cooling Cycle

condenser and adsorption beds and this is
written as:

Qchill + Qdes + Wcomp = Qgc + Qads

(17)

3.11 System Performance
After achieving a cyclic-steady-state, thesystem performance is calculated for one complete cycle time that is: adsorption time +pre-heating time + desorption time +pre-cooling time. The average chilling capacity
and COP during one cycle are expressed by eqs.
18 and 19, respectively.

∫ (m& c ) (T=t cycleQchill0chill , in

− Tchill , out )dt

tcycleCOP =4.

p chill

Qchill
Wcomp

(18)
(19)

Simulation Procedure

The simulation of the CO2 based hybrid
compression-adsorption cooling cycle transientmathematical model is carried out with theprogramming language Matlab. The differential equations are solved numerically with thefunction ODE45 and the thermophysicalproperties of the refrigerant are obtained bycalling the Refprop routines of NIST software.
The program starts with the initialization ofisotherm parameters and heat exchangers
constants (overall heat transfer coefficients,
heating capacities,…)) and these are furnishedin table 2. Then, the program initializes thevariables of the system (inlet temperatures,
adsorbing/desorbing time, desorbing pres-Tbed1 Tbed2 Tbed3 Tbed4

Thot,inTgc

Tcool,inTchill,in

sure,…). After that, the program estimates theinitial values of the differential equations parameters (Tevap, Tcond, Tbed1,..,4, c1,..,4). As shownin table 1, there are 8 modes during one cycle
time. The differential equations of each mode(energy balances and kinetic equations) are
put together in a function called by the main
program. The last results of the solved function are inputted as initial values to the next
mode’s differential equations. The key systemvariables are constantly computed and updated by numerical integration with time.

5.

Results and Discussion

The temperature profiles of the different
heat exchangers (gas cooler, evaporator andadsorber/desorber beds) are shown in Fig. 2.
The inlet temperatures of hot, cooling andchilled waters are taken as 85, 30 and 15 °C,
respectively. The adsorption and desorption
processes time are fixed to 600 s and theswitching time for pre-cooling and pre-heatingphases is set automatically to the time atwhich the pressure in the heated bed becomesequal to the gas cooler pressure. It is foundthat the compression-adsorption hybrid chiller
achieves cyclic steady-state operation within 3cycle-times. From Fig. 2, it is noticed that theevaporator temperature is fluctuating due tothe intermittent working of the thermal compression process. Its average value is equal to
9.7 °C for a chilled water inlet temperature of
15 °C. Moreover, the evaporator temperature
increases of about 2 °C during switching time,that is one bed switch from adsorbing process
to pre-heating process.
The outlet temperatures of chilled, cooling,and hot water as well as the compressor outlet

Tevap

Fig.2

Temperatures change of evaporator, gas cooler
and the 4 beds of the simulated cycle.

Fig.3

Outlet temperature profiles during half cycle time.

平成22年度

九州大学大学院総合理工学報告 第32巻 第3号

temperature are presented in Fig. 3. It is foundthat the cycle average chilled water outlet
temperature is 12.4 °C for a chilled water flowrate of 0.2 kg·s-1. The temperature approach
between the water inlet temperatures andwater outlet temperatures, at the end of adsorption and desorption processes, are found tobe 0.65 and 0.9 °C for hot and cooling sources,
respectively.
P-T-W (PresFig.
4
denotesthesure-Temperature-Uptake) diagram of the simulated thermal compression process. From
this figure, it is demonstrated that the hypothesis assumed in the mathematical model arevalidated. In fact, the desorption phase is realized at constant gas cooler pressure and theTable 2 Constants parameters of the cycle
Symbols

Value

Unit

C0

3.05

kg·kg-1

b0

0.11

kPa-1

Q

20.5

kJ·mol-1

t

0.663

-

Adsorption isotherms constants

System design constants

MMX3/bed(m& c )
p

lh t

p

hill

20

kg

2926

W K-1

836

W K-1

42100

J K-1

(Mcp)gc

2484

J K-1

(Mcp)evap

2484

J K-1

UAbed

1050

W K-1

UAgc

742

W K-1

UAevap850

W K-1(m& c )(Mcp)bed

16

adsorption phase pressure follows the saturation pressure dictated by the evaporator’s
temperature. The pre-cooling and pre-heatingphases are isosteric (parallel to dashed lines).
The P-h diagram of the hybrid system illustrates the energy saving by the thermal compression comparing to CO2 conventional vapor
compression cooling system (Fig. 5). It can benoticed that the gas cooler pressure and thegas cooler outlet temperature (point 4 in P-hdiagram) will have a big impact on the performance of the system. In fact, these two key
parameters dictate the amount of refrigerant
being evaporated. This is illustrated in Figs. 6
and 7 which show the change of cooling capacity (CC) and COP for different gas coolerpressures (Pgc) and for refrigerant outlet temperatures from gas cooler (Tgc,out) of 35 and
40 °C. Moreover, the performance of the CO2
based hybrid cycle (cycle 1-2-3-4-5) are compared to the conventional CO2 vapor compression cycle (cycle 1-2’-4-5) for the same evaporation temperature of 10 °C.
The specific cooling capacity (qevap = h1 - h5)of the hybrid system is similar to that of theconventional vapor compression system(dashed lines of Fig. 6). It increases with theincrease of Pgc as the fraction of liquid refrigerant being expanded is higher. However, therefrigerant flow rate of the hybrid cycle decreases with higher desorption pressures as
the amount desorbed decreases. This affects
the CC of the hybrid system which attains amaximum value above which the added capacity no longer compensates the decrease in refrigerant flow rate. The optimal cooling capacity values, 2.34 and 1.68 kW, are attained atdesorption pressures of 9.25 and 9.75 MPa for
Tgc,out of 35 and 40 °C, respectively.

Temperatures (0C)

Gas cooler 3

2'
Energy Saving
Thermalcompression

2
Mechanicalcompression

Fig.4

1

1k
g/k

P-T-W diagram of Maxsorb III-CO2 thermal compression process.

Evaporator

g

5

C=

kg
/kg

1. 5
C=

Sa
tur
ati
on

pr
ess
ure

Pr
e-

Pr
eco
ol

in
g

he
ati

ng

4

Fig.5

P-h diagram of CO2 based hybrid cooling cycle.

17

Performance Investigation of a Novel CO2 Compression-Adsorption Based Hybrid Cooling Cycle

160
140

Tgc,out = 308.15 K

2

120
100

1.5

80

Tgc,out = 313.15 K

1

60
40

0.5

20
0

0
7

Fig.6

Specific cooling capacity (kJ/kg)

Cooling capacity (kW)

2.5

8
9
10
Gas cooler pressure (MPa)11NomenclatureSpecific cooling capacity (dashed lines) and cooling
capacity (lines) of the hybrid cycle.

12
Hybrid cycle
Vapor comp. cycle

10

COP

8
6

Tgc,out = 308.15 K

42
Tgc,out = 313.15 K

0
7

8
9
10
Gas cooler pressure (MPa)tomobile engine is utilized to power the thermal compression cycle. The analyses presented
herein are based on the experimentally confirmed adsorption isotherm, adsorption kinetics, heat transfer coefficients and refrigerant
property routines. This novel compression-adsorption system achieves 2 to 3 times
higher COP comparing to that of the conventional vapor-compression systems and is suitable for automobile air-conditioning applications.

11

m&

A
b0
C
c
C0
CC
COP
cp
h
ksav
M,m
P
Q
Qst
R
T
t
U
W

Mass flow rate (kg s-1)
Area (m2)
Affinity at infinite temperature (Pa-1)
Adsorbed amount (kg kg-1)
Instantaneous uptake (kg kg-1)
Maximum uptake (kg kg-1)
Cooling capacity (kW)Coefficient of performance (-)
Specific heat capacity (J kg-1 K-1)
Enthalpy (J·kg-1)
Overall mass transfer coefficient (s-1)
Mass (kg)Pressure (Pa)Heat (W); Heat of adsorption (J kg-1)
Heat of adsorption (J kg-1)
Gas constant (J kg-1 K-1)
Temperature (K)Heterogeneity factor (-), time (s)Overall heat transfer coefficient (W m2 K-1)
Work (W)

Subscripts
Fig.7

Comparison of COP of hybrid cycle (lines) andvapor compression cycle (dashed line).

The COP rises with the increase of dischargepressure and attains two to three times higher
values comparing to that of the vapor compression cycle (Fig. 7). As explained above, therefrigerant flow rate of the hybrid cycle decreases for higher desorption pressures andtherefore the compressor work also decreases.
At the optimum CC reached, that is 2.34 and1.68 kW for Tgc,out of 35 and 40 °C, the COP
attains 9.5 and 8.6, respectively.

6.

Conclusions

An innovative environmentally benign aswell as high performance hybrid cooling system is presented. The system uses CO2 as therefrigerant and two steps compression; mechanical compression at the low pressure sideand thermal compression employing adsorption/desorption processes in the high pressure
side. In this system, the waste heat from au-

ads
bed
chill
comp
cool
cw
cycle
des
evap
gc
hot
in
MX3
out

Adsorption
Sorption heat exchanger
Chilled water
Compressor
Cooling water
Cooling water of gas cooler
Cycle time (s)Desorption
Evaporator
Gas cooler
Hot water
Inlet
Maxsorb III
OutletReferences
1) Kim, M.-H., Pettersen, J., and Bullard, C.W.,“Fundamental process and system design issues in
CO2 vapor compression systems”, Progress in Energy
and Combustion Science, Vol.30 (2004), pp.119-174.
2) Groll, E.A., and Kim, J.-H., “Review of recent advances toward transcritical CO2 cycle technology”,
HVAC and R Research, Vol.13 (2007), pp.499-520.
3) Hrnjak, P., “Thermodynamic possibilities and technological opportunities for improving the science ofrefrigeration in search for low global warming thermal systems”, Proc. Int. Symp. On Next-Generation
Air Conditioning and Refrigeration Technology, Tokyo(2010), K09.

平成22年度

九州大学大学院総合理工学報告 第32巻 第3号

4) Jribi, S., El-Sharkawy, I.I., Koyama, S., Saha, B.B.,“Study on activated carbon-CO2 pair: adsorption
characteristics and cycle performance”, Proc. Int.
Symp. On Next-Generation Air Conditioning and Refrigeration Technology, Tokyo (2010), P19.
5) Shen, C., Grande, C.A., Li, P., Yu, J., and Rodrigues,
A.E., “Adsorption equilibria and kinetics of CO2 and18

N2 on activated carbon beads”, Chem. Eng. J., Vol.160(2010), pp.398-407.
6) Pettersen, J., Hafner, A., Skaugen, G., and Rekstad,H., “Development of compact heat exchangers for CO2
air-conditioning systems”, Int. J. Refrigeration, Vol.21(1998), pp.180-193.