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Proceedings of ISES Solar World Congress 2007: Solar Energy and Human Settlement

942

hdh

⎛⎞

=++

⎜⎟

⎝⎠

˄2˅

h = 1326.96 W/(m

gk)

According to (3), the heat transfer on the evaporate side is:

=hA(t

−t

) ˄3˅

= 6.704 kW

3.2 Nature Heat Tranfer

It is unavoidable that some amount of heat transfer between

the Evaporator and Absorption side in the form vapor con-

duction, convection and radiation as a result of difference

in temperature.

q = q

+ q

˄4˅

In which, heat-conducting amount by vapor q

could be

known by (2-5)˖

qk t

=Δ ˄5˅

Which is: q

=2.16W

and the quantity of heat convection by vapor between the

two side is found by the theory of natural convection

heat transfer in limited space as a vertical interlayer,

which is

= 4.42 W

The radiation amount q

is calculated by the grid scheme,

and entirely thermal resistance could be found by (6), when

is (7).

FF F

εε

εϕε

−−

=+ +

˄6˅

()TT

−

˄7˅

What could be got is:

= 7.615W

So the total Nature Heat Tranfer amount is sum of the three ,

which is:

= 14.195W

3.3 Refrigerating Capacity

It can be indicated from the configuration of Evaporator-

Absorption that thermal equilibrium equation is:

= q

− q

˄8˅

where q

is the amount of heat transfer of falling-film

evaporation, q

is nature heat transfer quantity, q

is the

refrigerating capacity, based on previous results, which can

be calculated as follow:

=6.69 kW

The evaporated quantity of distilled water and the outlet

temperature of chilling water are computed by (9) and (10).

˄9˅

=−

˄10˅

Calculating result is m

= 2.698 g/s and t

=5.6ć,

respectively.

4. TEST DATA ANALYSIS

Some tests are taken under certain condition that how the

temperature of certain points varies along time. The data

analysis is as Fig. 3 to Fig. 5.

4.1 The Inlet And Outlet Temperature Of Hot Water

It is easy to know from Fig 3, the difference in inlet and out

3 SOLAR COLLECTOR TECHNOLOGIES AND SYSTEMS

943

let temperatures keeps at about 5ć, which meets the

theoretical value. In nearly 90 minutes, the inlet

temperature fluctuates within a narrow range, because the

hot water form outlet flows back to the hot water tank,

which makes the non-uniform temperature distribution in

the tank. The outlet temperature is affected by several

factors, such as the inlet temperature, running condition,

heat transfer inside the unit and so on.

Fig. 3: The inlet and outlet temperature of hot water curves

along time.

Fig. 4: The inlet and outlet temperature of cooling water

curves along time.

4.2 The Inlet And Outlet Temperature Of Cooling Water

The inlet and outlet temperature of cooling water becomes

higher as time goes on. The outlet cooling water flows back

to the tank, mixing with the water in it, which makes the

temperature becomes higher and distributes unevenly. At

the same, the water that has become warmer was not

replaced in time, which makes the inlet temperature of

cooling water goes higher.

4.3 The Inlet and Outlet Temperature of Chilling Water

According to Fig.5, the difference in inlet and out

let

temperatures keeps at about 3ć, which is not exactly

correspond to theoretical value. By analysis and checking

the unit, it can be explained as flows. Compared with the

original design, the inside structure is not very accurate and

the spray pipe and chilling water pipe are not fitted facing

to each other exactly. Only a part of distilled water sprays

on the chilling water pipe, which is not cooled fully.

Fig. 5: The inlet and outlet temperature of chilling water

curves along time.

5. CONCLUSIONS

In this paper, a compact absorption solar air-conditioner

system is introduced and its run mode is explained. On the

basis of theoretical arithmetic, the comprehensive heat

transfer coefficient on the Evaporator Side is calculated to

be 1326.96 W/(m

gk), while the distilled water absorption

ratio is 2.698g/s. The outlet temperature of chilling water is

5.6ć, and the theoretical Refrigerating Capacity is 6.69

kW, which is significant for optimal design and

improvement. By experimental study, some useful

performance curves are got. The temperature difference

between inlet and outlet chilling water is about 3 ć. Under

this operation condition, the refrigerating capacity of the

system can be got to be 2.14kW.

Proceedings of ISES Solar World Congress 2007: Solar Energy and Human Settlement

944

6. ACKNOWLEDGEMENT

This work was supported by the Natural Scientific

Foundation of China (No. 50576004) and the Basic

Research Foundation of Guangxi Province (No.0639034 ).

7. REFERENCES

(1) Conglin, “Performance Study of a Compact Solar Absorption

Air-Conditioner System”,[Master's Thesis], Beijing

Institute of Technology, 2007

(2) Hongfei Zheng, Yuyuan Wu, Hua Jiang, et al. 2003,

The study of the thermosiphon solution on elevation

tubes with lunate channel in solar absorption chiller[J],

Acta Energiae Solaris Sinica. 24(4): 466-47

(3) Lianying Zhang, Yuyuan Wu, Hongfei. “An experiment

investigation on performance of bubble pump with

lunate channel for absorption refrigeration system”. Int.

J. of Refrigeration, 2006, 29(5): 815-822

MODELLING AND TRANSIENT SIMULATION OF SOLAR-GROUND SOURCE

HEAT PUMP HEATING SYSTEM

Han Zongwei

School of Municipal and Environmental Engineering, Harbin

Institute of Technology, P.O. Box 235

Harbin 150090, China

taiyangshen1122@163.com

Zheng Maoyu

School of Municipal and Environmental Engineering, Harbin

Institute of Technology, P.O. Box 2644

Harbin 150090, China

zhengmaoyu@163.com

ABSTRACT

Solar - ground source heat pump (SGSHP) heating system

with phase change storage tank (PCST) is investigated. The

mathematical model of the system is developed, and the

transient numerical simulation is carried out in terms of this

model. The operation characteristic of the heating system is

analyzed during the heating period in Harbin. From the

results of the simulation, the average coefficient of

heating performance (COHP) of the heating system is

3.28 in heating period. In the initial and latter heating

period, the COHP is higher, because the system can be

operated without heat pump. During the middle heating

period the COHP and the operation stability of the system

are improved due to solar energy and soil alternately or

together as the heat source of heat pump. PCST is a very

important role in operation of the system. The system can

be operated more flexibly, effectively, and stably by the

charge and discharge heat of PCST, and the effect

becomes especially obvious in the initial and latter heating

period.

1. INTRODUCTION

Solar - ground source heat pump (SGSHP) heating system

have received much attention in recent years because of

the energy saving potential and environmental protection

when the systems are used in heating and cooling

applications [1].

There is some research on design, modeling, simulation,

experiments and optimization of the SAGSHP for heating

and cooling. From 1978−1981, some experimental and

theoretical researches on SAGSHP were performed by

Brookhaven National Laboratory (BNL, USA) [2].They

studied some coupling forms, that is, the use of the earth as

heat source/sink or storage element. Elasfouri and Hawas

developed a simplified model for simulating solar thermal

systems [3]. Krakow and Lin developed a computer model

for the simulation of multiple source heat pump

performance. The model was based on experimental

investigations of a solar-source heat pump and a

multiple-source heat pump system for cold climates [4].

Comakli et al. developed a dynamic simulation program for

a solar-assisted heat pump system with energy storage. [5].

Experimental performance study and exergoeconomic

analysis of a solar assisted ground source heat pump system

for greenhouse heating were done by Ozgener et al. [6]. All

above research provided the comprehensive foundation for

the design and application of SAGSHP heating system.

2. THE WORKING PRINCIPLE OF THE SYSTEM

2.1 The Component of the System

Fig. 1 shows that the system consists of five parts, which

are flat plate solar collector, PCST, vertical U tube soil

exchanger, heat pump unit, and fan coil.

Proceedings of ISES Solar World Congress 2007: Solar Energy and Human Settlement

946

Fig. 1: Schematic representation of the system.

2.2 The operation Mode of the System

For furthest utilizing solar energy and increasing the COP

of the system in heating period, the system should turn on

the corresponding operation mode when the outdoor

weather conditions change. The system has eight operation

modes:

 Mode1. Solar energy directly heats room without heat

pump and excessive heat is stored in PCST by

exchanger.

SC ⇒ PCST ⇒ FC

 Mode 2. PCST directly heats room without heat pump.

PCST ⇒ FC

 Mode 3. Solar energy is used as heat source of heat

pump to heat room, simultaneously charges and

discharges for PCST.

SC ⇒ PCST ⇒ HP ⇒ FC

 Mode 4. PCST is used as heat source of heat pump to

heat room.

PCST ⇒ HP ⇒ FC

 Mode 5. The ground-source heat pump (GSHP) heats

room.

SHE ⇒ HP ⇒ FC

 Mode 6. Solar energy and soil together are used as heat

source of heat pump to heat room.

SHE ⇒ SC ⇒ HP ⇒ FC

 Mode 7. Stop heating, and the heat from solar collector

is stored in PCST.

SC ⇒ PCST

 Mode 8. Stop heating, and the heat from solar collector

is stored in the soil around the soil

exchanger.

SC ⇒ SHE

3. MATHEMATICAL MODEL OF THE SYSTEM

The SGSHP heating system with PCST is a more

complicated system, which consists of several subsystems.

During development of the thermodynamic model, some

assumptions are made to simplify the numerical solution.

These assumptions can be summarized as follows:

(1) The temperature of PCST is homogeneous, and the

influences of stratification of temperature on the charge and

discharge heat process are ignored.

(2) The phase change material (PCM) is homogeneous and

isotropic, and its thermophysical properties of the PCM are

different for the solid and liquid phases but are independent

of temperature.

(3) The PCM behaves ideally, i.e. such phenomena as

property degradation and supercooling are not accounted

for.

(4) The PCM is assumed to have a definite melting point

(isothermal phase change).

(5) The vertical U tube exchanger is simplified as vertical

single tube, whose equivalent diameter is D

˙ 2 DǄ

(6) Comparing with the heat conduction in soil, the mass

transfer in earth is negligible; the heat transfer from earth to

soil exchanger is a pure heat condition process.

3.1 Solar Collector

When the collector turns on, the energy balance for the

solar collector can be expressed as:

()[()()]

u f Pf sco sci R sc sc e L sci am

QmCt t FAI Ut tτα=−= −−

(1)

When the collector turns off, the useful energy obtain by

the collector Q

˙0ˈthe temperature of absorber represents

the temperature of whole collector, and the energy balance

equation of solar collector is:

() ( )

scp

sc sc sc e sc L scp w

GAIAUtt

τα

=−−

(2)

3.2 Phase Change Storage Tank

The principle of the heat exchange in water tank is given by

exchanger

⎯

⎯⎯⎯⎯→

←⎯⎯⎯⎯⎯

water

⎯

⎯⎯⎯⎯→

←⎯⎯⎯⎯⎯

PCM. The

mathematical description of PCST can be written as

respectively:

3 SOLAR COLLECTOR TECHNOLOGIES AND SYSTEMS

947

Heat-transfer fluid (HTF) in exchanger

2()

Pf f he f w

Cm rUtt

π⋅=− −

(3)

Water in tank

( )() ()( )

GmCttUAtUAtt

dτ

=−−Δ−−i

(4)

PCM in tank

()()

div gradt UA t

τρ

∂

=+Δ

∂

(5)

Enthalpy H includes the two parts in Eq. (6), which are

sensible and latent heat

HCdtLf=+

∫

ii (6)

3.3 The Vertical U Tube Soil Exchanger

The controllable equation of HTF in exchanger is given by

fPf eq

ρπ

∂∂

(7)

Based on the above assumption, the controllable equation

of soil around the exchanger takes the following form

ss s

sPs

tt t

xxrr r

ρλλ

∂∂ ∂

∂∂

⎛⎞ ⎛ ⎞

⎜⎟ ⎜ ⎟

∂∂ ∂ ∂ ∂

⎝⎠ ⎝ ⎠

(8)

3.4 Heat Pump Unit and Fan Coil

To simplify the numerical solution, the characteristic of

heat pump unit are expressed by the fitted equations from

the manufacture’s catalog data, the heating capacity and

consumption of electricity are given by respectively

111hp ei ci

Qatbtc=−+ (9)

222

hp ei ci

at bt c=−+ (10)

To calculate heat supply of fan coil, the following

expression is found by fitting experimental date.

333fc fci i

Qatbtc=−+ (11)

3.5 The Thermodynamics Model for Heating Building

The transient heat load for building is evaluated by using

the equation

nam

nd od

−

(12)

The indoor temperature variation can be computed by

Eq.(13)

rm fc L

GQQ

dτ

=−

(13)

4. NUMERICAL SIMULATION OF THE SYSTEM

The simulation process mainly includes two stages. The

stage 1 is from October 16 to April 15 (the heating period

in Harbin). The stage 2 is from April 16 to October 15. The

main simulation parameters are shown as Table 1.

TABLE 1: BASIC PARAMETERS OF SIMULATION

Simulation area: Harbin (lat.N45.75 ;long.E126.77 )

Design heat load Q

10kW

Outdoor design temperature t

-26ć

Indoor design temperature t

18ć

Area of solar collector A

30m

Water volume in tank V

0.45m

PCM type CaCl

•6H

PCM total volume V

0.5m

PCM melting point T

29.9

PCM melting latent heat L 187.49kJ/kg

The number of soil exchanger n

she

The depth of soil exchanger 60m

Proceedings of ISES Solar World Congress 2007: Solar Energy and Human Settlement

948

The meteorological data of typical year in Harbin are

provided by national meteorological information centre of

China meteorological administration. Solar radiation and

outdoor temperature are shown as Fig. 2.

0 30 60 90 120 150 180 210 240 270 300 330 360

-25

-20

-15

-10

-5

The second stage

Temperature (ć)

Date

The first stage

Fig. 2(a): Curve of daily average outdoor temperature.

From Fig. 2 it is found distinctly that there is an

inconsistency between solar radiation and heat load of

building on time distribution, namely, solar radiation is

stronger in initial and latter heating period, but heat load

is smaller; heat load is larger in middle heating period, but

solar radiation is weaker. To ensure the effect of heating,

the design of system should be in terms of the weather

parameters in the middle heating period, which will lead

to the relative overmuch in initial and latter heating

period.

0 30 60 90 120 150 180 210 240 270 300 330 360

The second stage

Solar Radiation (MJ/m )

Date

The first stage

Fig. 2(b): Curve of daily solar radiation.

From Fig. 3, it is clear that the average temperature of

PCST is high than the melting temperature of PCM in the

initial and latter heating period. Solar energy is stored in

tank as sensible and latten heat. Due to the mostly heat

stored as latten heat in PCM, the temperature of tank is not

too higher, and the efficiency of solar collector and the

stored heat of tank are greatly increased. In the middle

heating period, the temperature is lower, and the

charge/discharge of the tank with only sensible heat.

0 30 60 90 120 150 180

Temperatureć

Date

Fig. 3: Curve of daily average temperature of the tank in

the first stage.

As shown in Fig. 4, the heat extracted is relatively smaller

in the initial and latter heating periods. Even the heat

extracted is negative at some day. In this stage the main

heat source of heating system is solar energy. With the solar

radiation gradually decreasing and the heat load increasing,

the heat extracted from soil correspondingly increases. In

middle heating period soil will be the main heat source.

0306090120150180

-1x10

1x10

2x10

3x10

Amount of heat (J/d)

Date

Fig. 4: Curve of daily total heat extracted from soil by

U-tube soil exchangers in the first stage.

3 SOLAR COLLECTOR TECHNOLOGIES AND SYSTEMS

949

From Fig. 5 we can see that in the initial and latter heating

period, the supplying heat is smaller, and the average

indoor temperature is relatively higher. In the middle

heating period, the supplying heat becomes large. But now

the average indoor temperature is lower, the lowest

temperature is 17.63ć, which is due to the discontinuously

controllable mode of the system and the heating deficiency

of GSHP heating system.

0 30 60 90 120 150 180

the daily average indoor temperature

the daily total supplied heat

Date

Temperature (ć)

-1x10

1x10

2x10

3x10

4x10

5x10

6x10

7x10

8x10

Supply heat (J)

Fig. 5: Curve of daily supplied heat and average indoor

temperature in the first stage.

In Harbin, the heating period is 183days. From the

simulation results we can know that the operation time of

this system is 3217 hours, where the heating condition is

2360 hours and the heat storage condition is 857 hours. The

total supplying heat is 94.83GJ and the total consumption

of electricity is 28.94GJ, and the average COP of the

heating system is 3.28. The characteristics of each mode are

shown as Table 2.

TABLE 2: THE OPERATION CHARACTERISTICS OF

EACH MODE

OPERATION

TIME (h)

SUPPLIED/

STORED

HEAT (GJ)

COHP

MODE 1 146 2.76 10.5

MODE 2 268 4.77 9.89

MODE 3 167 14.38 4.86

MODE 4 226 18.10 4.67

MODE 5 1017 35.52 2.51

MODE 6 486 19.29 3.00

MODE 7 692 18.82 —

MODE 8 165 15.50 —

6. NOMENCLATURES

A area (m

)

specific heat [J/(kg

⋅

K)]

D diameter (m)

heat removal factor

thermal capacity (J/K)

I solar radiation (W/m

)

L latent heat of PCM (J/kg)

P consumption of electricity (W)

Q heat power (W)

U heat-transfer coefficient [W/(m

⋅

K)]

liquid fraction

m mass flow rate (kg/s)

q heat exchange capacity per unit length (W/m)

t temperature (ć)

r radial axis (m)

x axial axis (m)

Greece letter

τ time (s)

λ thermal conductivity [W⁄(m

⋅

K)]

ρ Density (kg/m

)

υ velocity (m/s)

subscript

am ambient c condenser

d design value e evaporator

eh exchanger f

heat transfer

fluid

fc fan coil i input

L heat load n indoor

o output or outdoor p absorber

rm heating room s soil

sc solar collector t PCST

u useful w water

7. REFERENCES

(1) Kavanaugh SP, “Field test of vertical ground-coupled

heat pump in Alabama”, ASHRAE Trans, 1992,

98(2):p.607-616

(2) Metz P D, “the Use of Ground Coupled Thanks in Solar

Proceedings of ISES Solar World Congress 2007: Solar Energy and Human Settlement

950

Assisted Heat Pump System”, Transaction of ASME,

Journal of Solar Energy Engineering, 1982,

104(4):p366-372

(3) Elasfouri AS, Hawas MM, “A simplified model for

simulating solar thermal systems”, Energy Convers

Mgmt, 1987, 27:p1-10

(4) Krakow KI, Lin S, “A computer model for the

simulation of multiple source heat pump performance”,

ASHRAE Trans, 1983, 89:p590-616

(5) Comakli Ő, Kaygusuz K, Ayhan T, Arslan F,

“Experimental investigation and a dynamic simulation

of the solar-assisted energy storaged heat pump

system”, Solar Energy, 1993, 51:p147-158

(6) Ozgener O, Hepbasli A, “Exergoeconomic Analysis of a

Solar Assisted Ground-Source Heat Pump Greenhouse

Heating System”, Appl Thermal Eng, 2005,

25:p1459-1471

DISTRIBUTED AND STEADY MODELING OF THE PV EVAPORATOR

IN A PV/T SOLAR ASSISTED HEAT PUMP

Jie Ji, Hanfeng He, Wei He, Gang Pei, Keliang Liu

Dept. of Thermal Science and Energy Engineering,

Uni. of Science and Technology of China

Hefei 230026, China

Fax.86-551-3601652, Email:jijie@ustc.edu.cn

ABSTRACT

A specially designed direct-expansion evaporator (PV

evaporator), which is laminated with PV cells on the front

surface is adopted in a photovoltaic/thermal solar assisted

heat pump (PV/T SAHP) to obtain both thermal energy and

electricity from solar radiation. A distributed and steady

model is presented to describe the performance of the PV

evaporator. In the model, the influence of the pressure drop

on the refrigerant properties, such as evaporating

temperature, density, enthalpy, is taken into account. The

model is capable of predicting the spatial distribution of

pressure, temperature, vapor quality, void fraction and

enthalpy of the refrigerant. Two-dimensional distribution of

the evaporator temperature and PV efficiency are also

given by the model.

1. INTRODUCTION

Well know as a non-polluting, inexhaustible, and clean

energy source, solar energy has received considerable

attention in recent decades. The solar-assisted heat pump

technology and photovoltaic technology are two different

means of utilization of solar energy.

It is a well-known fact that the electrical efficiency of the

PV cells drops with the increase in operating temperature.

A PV/T system which adopts water or air as coolant of the

PV cells can overcome the limitation by bringing down its

operating temperature. The best design is to re-use the heat

energy removed by the coolant. Russell and Kern [1]

presented an indirect-expansion PV/T solar assisted heat

pump system which combines the above two solar

technologies into a single system in 1979. A water-type

PV/T collector was employed in the system for duel

production of electricity and thermal energy. They

investigated the system with TRANSYS for residence in

New York and Fort Worth climates. Analysis of the

technical and economic results was also discussed in their

research. Ito et al. [2] coupled the photovoltaic technology

with a direct-expansion solar assisted heat pump system. A

PV/T solar collector with refrigerant as coolant was

employed in their system to gain both thermal energy and

electricity from solar radiation. The experimental results

showed that the PV cells had weak influence on the thermal

performance. Hence their research mainly focused on the

thermal performance of the system. The results showed that

the coefficient of the performance (COP) of the heat pump

system could reach 6.0 when the temperature of the cooling

water in the condenser was maintained at 40 .ć

A specially designed tube-sheet direct-expansion

evaporator, which is laminated with PV cells on the front

surface is adopted in a photovoltaic/thermal solar assisted

heat pump (PV/T SAHP). To enlarge the contact area of the

copper coil and the absorbed panel, the copper coil is

located in the grooves which are formed by an aluminum

alloy panel and an aluminum panel as shown in Fig. 1. A

fraction of the solar energy incident upon the PV

evaporator is converted to electricity by the PV cells and

most of the rest is converted to thermal energy which is

Труды Всемирного конгресса Международного общества солнечной энергии - 2007. Том 2

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