propane heat pump with low refrigerant charge: design and laboratory tests

Propane heat pump with low refrigerant charge:

design and laboratory tests

Primal Fernando*, Bjorn Palm, Per Lundqvist, Eric Granryd

Division of Applied Thermodynamics and Refrigeration, Department of Energy Technology,

Royal Institute of Technology (KTH), SE-100 44 Stockholm, Sweden

Received 12 January 2004; received in revised form 28 June 2004; accepted 30 June 2004

Abstract

Independently of the choice of refrigerant, environmental and or safety issues can be minimised by reducing the amount of

refrigerant charge per heat pump or refrigeration system. In the investigation reported here, a laboratory test rig was built,

simulating a water-to-water heat pump with a heating capacity of 5 kW. The system was designed to minimize the charge of

refrigerant mainly by use of mini-channel aluminium heat exchangers. It was shown that the system could be run with 200 g of

propane at typical Swedish operating conditions without reduction of the COP compared to a traditional design. Additional

charge reduction is possible by selecting proper compressor lubrication oils or by using a compressor with less lubrication oil.

q 2004 Elsevier Ltd and IIR. All rights reserved.

Keywords: Heat pump; Water-water; Refrigerant charge; Propane; Design; Experiment

Pompe a chaleur a propane et a faible charge:

conception et essais en laboratoire

Mots cles: Pompe a chaleur; Eau-eau; Charge en frigorigene; Propane; Conception; Experimentation

1. Introduction

Sweden has several relatively big (for European

standards) production units of heat pumps. The Swedish

heat pump market is also traditionally larger than other

European markets. The number of installed and sold heat

0140-7007/$35.00 q 2004 Elsevier Ltd and IIR. All rights reserved.

doi:10.1016/j.ijrefrig.2004.06.012

* Corresponding author. Tel.: C46-8-790-8941; fax: C46-8-20-

30-07.

E-mail addresses: [email protected] (P. Fernando), bpalm

@energy.kth.se (B. Palm), [email protected] (P. Lundqvist),

[email protected] (E. Granryd).

pumps annually in Sweden is at the same level as for the rest

of Europe. The main reason for this large interest is the cold

climate in combination with low price of electricity at the

consumer level, being, per kWh, comparable to the price of

oil. Several countries, however, foresee a strong increase in

sales of heat pumps within coming years. Swedish heat

pumps generally have good reputation for high quality.

Fig. 1 shows the number of heat pumps installed per year

in Sweden [1]. The number of installed heat pumps started

to increase rapidly in 1995. The number of heat pumps

represent in the figure in year 2003 is only up to the month of

August. The vast majority of them are installed in single-

family houses. The most common type is an indirect (brine

International Journal of Refrigeration 27 (2004) 761–773

www.elsevier.com/locate/ijrefrig

http://www.elsevier.com/locate/ijrefrig

Nomenclature

A heat transfer area (m2)

A-BH-25 average temperature of the borehole, 25 m deep

(8C)

A-BH-50 average temperature of the borehole, 50 m deep

(8C)

A-ut average outside temperature (8C)

COP coefficient of performance, heating

cp specific heat capacity (J kgK1 KK1)

C-170 charge 170 g

C-201 charge 201 g

C-240 charge 240 g

C-265 charge 265 g

GWP green house warming potential

h refrigerant enthalpy (J kgK1)

LMTD log mean temperature difference (K)

m mass flow rate (kg sK1)

M-BH-25 minimum temperature of the borehole, 25 m

deep (8C)

M-BH-50 minimum temperature of the borehole, 50 m

deep (8C)

M-ut minimum outside temperature (8C)

M-Br-in minimum brine inlet temperature (8C)

ODP ozone depletion potential

Q heat capacity (W)

SH super heat (K)

SC sub cool (K)

t temperature (8C)

U overall heat transfer coefficient (W mK2 KK1)

Subscripts

ave average

c condenser

c,in condenser inlet

c,out condenser outlet

c,dsh,r condenser de-superheat region

c,sc,r condenser sub-cool region

c,tp,r condenser two-phase region

c,sl condenser, saturated liquid

c,sg condenser, saturated gas

c,sc condenser, sub-cooled liquid

e evaporator

e,in evaporator inlet

e,out evaporator outlet

e,sh,r evaporator superheat region

e,tp,r evaporator two phase region

e,sg at evaporator saturated gas

e,sh at evaporator super heat

g glycol

g,out glycol outlet

g,in glycol inlet

g,sg glycol at refrigerant saturated gas

min minimum

r refrigerant

r,sh refrigerant, superheated gas

r,sg refrigerant, saturated gas

r,in,e refrigerant inlet in evaporator

r,in,c refrigerant, condenser inlet

r,sl refrigerant, saturated liquid

w water

w,out water outlet

w,in water inlet

w,sg water at condenser saturated gas

w,sl water at saturated liquid

P. Fernando et al. / International Journal of Refrigeration 27 (2004) 761–773762

loop) ground or rock-coupled heat pump, connected to the

water circulating heating system of the house. Usually, the

heat pump is chosen to cover more than 90% of the annual

heating need of a house and 50–60% of the heating demand

of the coldest day. The total heating demand covered by heat

pumps in Sweden is estimated to be about 15 TWh/year,

thus reducing the total energy demands by about 10 TWh/

year. Environmentally adopted and energy effective heat

pumps can play an important role as replacement of the

presently dominant energy technology, in Sweden, the

direct electric heating.

The ozone depletion potential (ODP) and green house

warming potential (GWP) of some commonly used

refrigerants have led to increased research efforts for

developing other refrigerants, which are environmentally

friendly. Propane is a refrigerant, which has no ODP and

extremely low (!20) GWP [2], compared to many

currently used refrigerants. Also thermodynamic and

transport properties are the same as or better than the most

popular refrigerants used in refrigeration systems. Propane

is not corrosive in combination with many materials such as

aluminium, brass, bronze, copper, stainless steel, silver, etc.

Therefore it is fully compatible with existing components

such as heat exchangers, expansion valves, compressors,

lubricants and copper tubing, which are commonly used in

current refrigeration systems [3]. Its major drawback,

however, is the flammability.

Independently of the chosen refrigerant it is of interest to

decrease the total charge of the system, if this can be done

without negative influence on the COP of the system. For

HFC and HCFC systems, reduced charge means less

environmental impact in case of leakage. For systems with

flammable refrigerants lower charge is important for safety

reasons. An ongoing project at The Royal Institute of

Technology (KTH), Stockholm, aims to increase the

existing knowledge about heat pumps and refrigeration

systems designed for the lowest possible refrigerant charge,

with the goal of designing a heat pump system with a charge

of less than 150 g (with propane) having a heating capacity

of the order of 5 kW or more. This should be reached

Fig. 1. The number of heat pumps installed in Sweden from 1986 to August 2003 [1].

P. Fernando et al. / International Journal of Refrigeration 27 (2004) 761–773 763

without reduction of COP. The level of the heating capacity

was chosen taking into consideration the Swedish and

eventually cumulative European markets for single-family

house heat pumps.

2. Charge minimisation with narrow channel heat

exchangers

2.1. Narrow channel heat exchangers

Heat exchangers are the main concern in charge

minimisation of a heat pump as the high interior volumes

of the heat exchangers can cause large refrigerant charge.

Plate type heat exchangers are widely used in liquid/liquid

applications in refrigeration and air conditioning due to their

compactness, high heat transfer coefficients and low

pressure drop. Compared to most other types, the internal

volume is low. However, if very low charges are desired the

amount of refrigerant contained inside during normal

operation is still too high [4]. Experimental investigations

show that using micro-channel heat exchangers in refriger-

ation systems can reduce substantially the refrigerant charge

[5]. It is obvious that the reduction of channel diameter lead

to increase in the heat transfer area per unit heat exchanger

volume. Previous tests done to compare the plate heat

exchangers with mini-channel aluminium heat exchangers

showed a considerable charge reduction when switching to

mini-channel heat exchangers in the same heat pump [6].

Especially, the outlet of the condenser and the inlet of the

evaporator (headers) of plate heat exchangers contain lots of

refrigerant and acts as a liquid pool. The headers of narrow

(micro or mini) channel heat exchangers can be designed to

avoid this collection of liquid refrigerant. In short, the use of

mini-channel heat exchangers leads to wide possibilities of

charge reduction.

Narrow channel tubes with various diameters and

lengths are manufactured for many industrial applications.

The most common materials are aluminium and copper,

which both have high thermal conductivities. New manu-

facturing methods allow producing tubes with small wall

thickness and even with internal longitudinal fins. This is a

great advantage when designing heat exchangers with high

heat transfer areas and low internal volumes. The heat

exchangers in heat pumps should be designed for small

temperature differences, which is economical due to long

operating hours. Through, narrow channels and rough or

finned surfaces often lead to higher pressure drops.

However, this can be avoided by using a sufficient number

of parallel channels.

2.2. Mini-channel aluminium heat exchangers

Two liquid-cooled heat exchangers have been fabricated

with multiport aluminium tubes. Fig. 2 shows the cross

section of the used tube. As shown, the tube had six

channels. The inner cross section of the four centre channels

was 1!2.65 mm2 and of the other two channels 1!1.45 mm2 with an additional end curvature with 1 mm in

diameter each, giving a hydraulic diameters of 1.45 mm for

the centre channels and 1.35 mm for the end channels. The

wall thickness of the tube was 0.5 mm and the length of each

tube was 651 mm (Figs. 3 and 4).

The evaporator and the condenser consisted of 30 and 36

aluminium tubes in two parallel rows, respectively. A shell

and 31 baffle plates were made for containing and directing

Fig. 2. The tube cross section.

Fig. 3. Inlet of the heat exchanger showing the tube rows.

Fig. 4. (a) A picture of the exchanger; view inside the shell. (b) A

picture of the heat exchanger.


the secondary refrigerant on the shell side of each heat

exchanger. For the calculations, it was assumed that the

baffle plates of the heat exchangers at as fins. The outer areas

(tube areas between baffle plates and areas of the baffle

plates) of the evaporator and the condenser were 0.822 and

0.985 m2, respectively. The two heat exchangers were

mounted in a test rig, simulating typical heat pump

conditions and the test results were compared with results

from corresponding tests with plate type heat exchangers.

The overall heat transfer coefficients of the aluminium mini

channel heat exchangers, working as evaporator and

condenser were 50 and 60% higher, respectively, than

those of the plate heat exchangers. The pressure drop of the

aluminium evaporator was 70% higher than that of the plate

type evaporator, but the pressure drops of both types of

condensers were negligibly small. The total internal

volumes of the aluminium heat exchangers were less than

50% of those of the plate heat exchangers. It was found that

the amount of refrigerant charge in the heat pump with the

plate type heat exchangers was about 300 g, and that with

mini channel heat exchangers about 200 g. The measured

amount of refrigerant in the plate type evaporator was 69 g

and for the mini channels aluminium evaporator 28 g. The

amount of refrigerant in the plate type condenser was 125 g

and for the mini channels aluminium condenser 96 g [7].

3. Experimental set-up heat pump

Fig. 5 shows a schematic diagram of the experimental

set-up. It consisted of a scroll compressor, mini-channel

aluminium heat exchangers, working as evaporator and

condenser and an electronic expansion valve. Measuring the

temperature and the pressure of the evaporator outlet

controlled the electronic expansion valve. The superheat

was calculated from these readings in a PID controller. The

output signal from the PID controller could also set

manually. Pneumatically operated ball valves were placed

at the inlet and outlet of the heat exchangers. The four ball

valves could be closed simultaneously within 0.25 s, in

order to trap the refrigerant in the different sections during

operation. The condenser was cooled by water and the heat

supply to the evaporator was arranged by an electrically

heated glycol solution. The water flow rate through the

condenser and the glycol flow rate through the evaporator

were measured with electronic flow meters. Pressure

measurements and pressure drop measurements were

taken by two pressure transducers connected at the

evaporator and condenser inlets and two differential

pressure transducers connected between inlet and outlet of

the heat exchangers. Thermo-couples were provided for the

temperature measurements. The thermo-couples were

calibrated by inserting them into an ice-bath (assuming the

temperature of the ice-bath 0 8C). The calibration curves and

the accuracies of the flow meters were taken as given by the

manufacturer. The flow meters were also tested with known

Fig. 5. Schematic diagram of the experimental heat pump.


flow rates, which confirmed the given accuracies. The

pressure transducers were calibrated with standard calibrat-

ing equipment. The measurements accuracies were; tem-

peratures measurements G0.04 at 0 8C, flow measurements

0.5% of measured value and pressure measurements G0.02% of the measured value. All the instruments were

calibrated several times during the test period.

Fig. 6. Temperature profile in the evaporator.

4. Calculations

The heat transfer rate of the condenser and the

evaporator can be calculated from the temperature change

of the secondary refrigerant, the mass flow rate and the

specific heat of the fluid

Qc Zmwcp;wðtw;out K tw;inÞ (1)

Qe Zmgcp;gðtg;in K tg;outÞ (2)

The heat transfer rate of the condenser and evaporator can

also be calculated from the enthalpy change of the

refrigerant and the refrigerant mass flow rate

Qc Zmrðhc;in Khc;outÞ (3)

Qe Zmrðhe;out Khe;inÞ (4)

The overall energy balance of the evaporator is

Qe ZUeAeLMTDe (5)

The logarithmic mean temperature difference (LMTD) of

the evaporator was calculated considering a division of the

evaporator into two parts: the evaporation part and the

superheating part [8]. Fig. 6 shows the temperature profile of

the evaporator.

The overall heat transfer coefficient of the evaporator is

defined as

Ue ZUe;tp;rAe;tp;r CUe;sh;rAe;sh;r

Ae

(6)

The overall energy balance of the two regions are

Qe;tp;r ZUe;tp;rAe;tp;rLMTDe;tp;r (7)

Qe;sh;r ZUe;sh;rAe;sh;rLMTDe;sh;r (8)

The Eqs. (5)–(8) can be summarised as

LMTDe ZQe

Qe;sh;r

LMTDe;sh;rC

Qe;tp;r

LMTDe;tp;r

(9)

The heat transfer from glycol to refrigerant in the

different regions can be calculated as

Qe;tp;r Zmrðhe;sg Khe;inÞ (10)

Qe;sh;r Zmrðhe;sh Khe;sgÞ (11)

The LMTDs of the two regions


LMTDe;sh;r Zðtg;in K tr;shÞK ðtg;sg K tr;sgÞ

lntg;inKtr;sh

tg;sgKtr;sg

� � (12)

LMTDe;tp;r Zðtg;sg K tr;sgÞK ðtg;out K tr;in;eÞ

lntg;sgKtr;sg

tg;outKtr;in;e

� � (13)

The overall energy balance of the condenser is

Qc ZUcAcLMTDc (14)

The LMTD of the condenser was calculated considering

division of the condenser into three parts [9]: de-super-

heating region, condensing region and sub-cooling region.

Fig. 7 shows the temperature profile of the condenser.

Fig. 7. Temperature profile in the condenser.

The overall heat transfer coefficient of the condenser is

defined as

Uc ZUc;sc;rAc;sc;r CUc;tp;rAc;tp;r CUc;dsh;rAc;dsh;r

Ac

(15)

The overall energy balance of the three regions

Qc;sc;r ZUc;sc;rAc;sc;rLMTDc;sc;r (16)

Qc;tp;r ZUc;tp;rAc;tp;rLMTDc;tp;r (17)

Qc;dsh;r ZUc;dsh;rAc;dsh;rLMTDc;dsh;r (18)

The Eqs. (14)–(18) can be summarised as,

LMTDc ZQc

Qc;dsh;r

LMTDc;dsh;rC

Qc;tp;r

LMTDc;tp;rC

Qc;sc;r

LMTDc;sc;r

(19)

The heat transfer from refrigerant to water in the different

regions of the condenser can be calculated as

Qc;sc;r Zmrðhc;sl Khc;scÞ (20)

Qc;tp;r Zmrðhc;sg Khc;slÞ (21)

Qc;dsh;r Zmrðhc;in Khc;sgÞ (22)

The LMTDs of different regions can be described as

LMTDc;dsh;r Zðtr;in;c K tw;outÞK ðtr;sg K tw;sgÞ

lntr;in;cKtw;out

tr;sgKtw;sg

� � (23)

LMTDc;tp;r Zðtr;sg K tw;sgÞK ðtr;sl K tw;slÞ

lntr;sgKtw;sg

tr;slKtw;sl

� � (24)

LMTDc;sc;r Zðtr;sl K tw;slÞK ðtc;sc K tw;inÞ

lntr;slKtw;sl

tc;scKtw;in

� � (25)

The pressure drop of the evaporator was considered in

calculations of the refrigerant side temperatures, but was not

considered in the condenser, since the pressure drop of the

condenser was negligibly small.

5. The performance of the heat pump with mini-channel

aluminium heat exchangers

5.1. Optimum charge tests

The heat pump was tested with selected heat source

temperatures and with a constant heat sink temperature. The

selected heat source temperatures were K10, K2, 6 and

12 8C, and the heat sink temperature was 40 8C. The cooling

water inlet temperature was kept at 33 8C, to obtain a 7 K

temperature increase of the water. The temperature change

of the glycol was kept at about 3 K for the heat source

temperatures of K10 and K2 8C and at about 3.75 and

4.4 K for the heat source temperatures of 6 and 12 8C,

respectively.

The refrigerant quantity (propane) in the system was

varied for each heat source and heat sink temperature

combination and the coefficient of performance (COP) of

the system was determined. The superheat at the evaporator

outlet was slightly increased with decreasing evaporation

temperature for lower charges in order to avoid gas bubbles

in the expansion valve inlet. This was done in order to

simulate the performance of a correctly sized thermostatic

expansion valve. The minimum (desired) superheat was

maintained around 5 K. The sub-cooling at the condenser

outlet was allowed to increase for higher charges. The heat

pump was tested over 1000 h for all heat source/sink

combinations.

Fig. 8 shows the COP as a function of the refrigerant

charge at different heat source temperatures. For a given

heat source temperature, the COP was more or less constant

when the charge was above a certain minimum level. Below

this level, the COP dropped significantly. The optimum

refrigerant charges resulting in the highest COP, for the four

tested heat source/sink combinations were170, 200, 240 and

265 g, in order from lowest to highest heat source

temperature.

Fig. 9 shows the variation of the heating capacity with

Fig. 8. COP of the condenser vs refrigerant charge for four heat source temperatures (K10, K2, 6, 12 8C) a constant heat sink temperature

40 8C.


the refrigerant charge. As for the COP the capacity is more

or less constant as long as the charge is above the optimum

charge. The capacity and COP reduction for lower

refrigerant charges is caused by the decrease of the

evaporation temperature. The reason for this decrease is

the ‘starvation’ of the evaporator, leading to an increase of

the superheat and to an outgoing refrigerant temperature

close to the inlet glycol temperature.

5.2. The overall heat transfer coefficients of the evaporator

and the condenser

Some refrigerant charges, around the optimum charge

for each test condition, were selected. The capacity variation

Fig. 9. Heat capacity of the condenser vs charge for four heat source tempe

40 8C.

and LMTD variation of the condenser and evaporator for the

selected charges were calculated and the results are plotted

in Figs. 10 and 11.

Fig. 10 shows the LMTD and condenser capacity (Qc)

variation with refrigerant charge (at different heat source

temperatures). The condensing temperature was increased

slightly with increasing heat source temperatures, mainly

due to the increased mass flow resulting more refrigerant

occupation in condenser. That was probably the reason why

higher LMTDs were recorded at higher heat source

temperatures, although the temperature change in the

water was kept about 7 K for all tests. For each heat source

temperature the LMTD increased slightly with increasing

charge due to the increasing of the sub-cool.

ratures (K10, K2, 6, 12 8C) and a constant heat sink temperature of

Fig. 10. Heat capacity and the LMTD of the condenser vs refrigerant charge in the system.

Fig. 11. Cooling capacity and LMTD of the evaporator vs refrigerant charge in the system.

Fig. 12. Variation of super heat and sub-cooling.


Fig. 13. Overall heat transfer coefficients of evaporator and condenser vs refrigerant charge.


Fig. 11 shows that the LMTD of the evaporator was in

the range from 4 to 5 K for all four tests conditions and that

the evaporator capacity was strongly influenced by the heat

source temperature. For each heat source temperature, the

charge had very little influence on the capacity or the LMTD

of the evaporator.

Fig. 12 shows the variations of the superheat at the

evaporator outlet and the sub-cooling at the condenser outlet

for different heat source temperatures and refrigerant

charges close to the optimum. The superheat was kept

between 4 and 6.5 K and the sub-cooling was depending on

the system charge, since no receiver was equipped to the

heat pump. For each heat source temperature the superheat

was independent of the refrigerant charge, while the sub-

cooling was strongly coupled to the charge. As shown, the

optimum charges for each heat source temperature corre-

spond to a sub-cooling of 4–5 K.

Fig. 13 shows the variation of the overall heat transfer

Table 1

Refrigerant charge distribution in various components

Heat source/heat sink

temperatures (8C)

K10.22/40.77 K2.09/40.67

Evaporation/condensing

temp. (8C)

K16.49/39.73 K8.79/40.44

Evaporator/condenser

capacity (kW)

2.67/4.00 3.66/5.01

Super heat/sub cool (K) 4.66/3.91 5.30/4.76

Optimum refrigerant

charge (g)

170 201

Measured in evaporator

(g)

27 23

Measured in condenser

(g)

69 80

Measured in liquid line

(g)

24 24

Measured in compressor

(g)

50 74

coefficients of the evaporator and the condenser for the

selected charges and heat source temperatures. The LMTD

of the evaporator was almost the same for all test conditions.

The overall heat transfer coefficient of the evaporator

depends significantly on the heat source temperature, which

influences the heat flux and mass flow of the refrigerant

through the evaporator. This is in agreement with the

following expectations; the boiling heat transfer coefficients

should increase with increasing mass flux, heat flux and

evaporation temperature. For the condenser, the overall heat

transfer coefficient was almost independent of the heat

source temperatures. This was due to the higher LMTDs at

higher heat source temperatures, although the capacity (Qc)

was higher at higher heat source temperatures. It can be

observed a decreasing in overall heat transfer coefficient at

the selected refrigerant charge ranges, since the increasing

of the refrigerant charge leads to a increasing the sub cool

resulting increasing of the LMTD.

6.93/40.77 12.64/40.74

K0.51/41.44 4.55/39.95

4.90/6.31 5.89/7.23

5.18/4.56 5.77/3.82

240 265

25 26

90 93

23 24

102 122


5.3. Charge distribution tests

The purpose of these tests was to determine the

distribution of the refrigerant charge in the different parts

of the system. For each heat source temperature tested, the

system was charged with the optimum refrigerant charge.

Measurements were taken at stable conditions for a given

heat source/sink temperature combinations and the four ball

valves of the heat pump were closed simultaneously while

the system was in operation. A small cylinder (one litre

capacity) was connected to the isolated sections by a flexible

hose and the cylinder was immersed in a larger container,

which was filled with liquid air. The cooling of the cylinder

caused a low pressure inside and thereby all the refrigerant

in different sections was drained to the cylinder. The

differences between original refrigerant charge and the total

refrigerant drained were less than 5 g in all tests. The

drained refrigerant was weighed. The accuracy of the

measurements was G1 g.

Table 1 shows the refrigerant charge distribution in the

heat pump for the four tested heat source temperatures for

the optimum charge. The amount of refrigerant in the

evaporator was almost same for all four tests, even though

the evaporation temperature varied from K16.5 to C4.5 8C.

Also, the amount of refrigerant in the liquid line was the

same in all four experiments. The amount of refrigerant in

the condenser varied from 69 to 93 g, showing an increase of

24 g in spite of the fact that the condensing temperature was

almost constant. The refrigerant amount in the compressor

was increased from 50 to 122 g, showing an increase of

72 g. The results of the charge distribution tests indicate that

a considerable amount of the refrigerant is located in the

compressor. It increases rapidly with increasing evaporation

temperature. This is a problem if, preferably, the system

should be able to run with the same charge at different

conditions. The lubrication oil type used in the compressor

was ester type oil (polyol ester). In parallel experiments, the

solubility of propane in different lubrication oils was tested.

In these experiments, the solubility of propane with two

polyol ester (POE) type oils and with one polyalkylene

glycol (PAG) type oil was tested. The solubility was tested

at different saturation temperatures and found that increased

at higher temperatures and pressures. The solubility of

propane in ester oil was higher than PAG oil [10].

Solubility of refrigerant in lubrication oil is positive

outside of the compressor as it aids the oil return to the

compressor and also keeps the heat transfer surfaces free of

oil. Within the compressor high solubility causes a decrease

in lubricant viscosity, which may be harmful for the proper

lubrication [11]. The solubility of refrigerant in lubricants

also depends on the viscosity of the lubricant. A lubricant

with higher viscosity has low solubility in a refrigerant

compared with the same lubricant type with lower viscosity

[12]. The selection of less soluble lubricant oil with propane

or lubricant oil with higher viscosity than that used for these

experiments, will allow further reduction of refrigerant

charge.

The pneumatically operated ball valves were placed

close to the heat exchangers (Fig. 5). The refrigerant

amounts measured in the heat exchangers were; the total

amounts in the heat transfer tubes, two end caps (headers)

and the tube lengths between ball valves and heat

exchangers. The total inside volume between the ball valves

surrounding the evaporator is 376 cm3; of this, the top end

cap with the tube length between the ball valve and the heat

exchanger has a volume of 52 cm3, the bottom end cap with

the tube length between the ball valve and the heat

exchanger has a volume 42 cm3 and heat transfer tube

volume is 279 cm3. The total inside volume between the

valves surrounding the condenser is 437 cm3; of which the

top end cap with the tube length between the ball valve and

the heat exchanger represents a volume of 53 cm3, the

bottom end cap with the tube length between the ball valve

and the heat exchanger has the volume of 49 cm3 and the

heat transfer tube volume is 335 cm3. The end caps and the

extra tube lengths occupy 25% of the total volume. Half of

this 25% volume is in the bottom side of the evaporator and

condenser where the refrigerant is at least partly in liquid

phase. Since both, the evaporator and the condenser were

placed vertically; a considerable amount of refrigerant was

in the evaporator bottom side end cap and condenser bottom

side end cap. A redesign of these parts should result in lower

liquid hold-up. The line between the two ball valves at the

condenser outlet and the evaporator inlet was considered as

the liquid line. The liquid line consisted of: a sight glass,

the expansion valve and some extra tubing to facilitate the

fitting of the ball valves. The refrigerant amounts in the

liquid line were consistent for all tests with a total mass of

about 24 g. The introduction of smaller ends caps and

exclusion of extended tube lengths for the ball valves may

further reduce the optimum charge at least by 20 g.

5.4. Performance of the heat pump for a given refrigerant

charge

In a separate test series, the system was charged with the

optimum refrigerant charges determined for the heat source

temperatures K10, K2, 6 and 12 8C. For each charge the

heat source temperature was varied from K10 to C12 8C,

while keeping the heat sink temperature constant at 40 8C.

Fig. 14 shows the influence of the heat source

temperature on the COP. The minimum sub-cooling of the

condenser was about 2 K and was increased for lower heat

source temperatures due to backup of refrigerant into the

condenser. The minimum super heat was kept at 4–6 K but

was allowed to increase at higher heat source temperatures

by closing the electronic expansion valve, thereby also

lowering the evaporation temperature. This was done in

order to prevent uncondensed gas to reach the expansion

valve inlet.

Fig. 14. Variations of COP with heat source temperature for different refrigerant charges.


6. Heat source temperature

It is obvious that of higher heat source temperatures are

more favourable in heat pumps. The ground heat source is

the dominant heat source for heat pumps in the range above

5 kW in Sweden. The most common method of extracting

heat from the ground is the circulation of brine in an 80–

150 m deep borehole. These boreholes provide excellent

working conditions and high reliability, but at relatively

high cost [13]. Figs. 15 and 16 show the monitored

temperature measurements in a single-family heat pump

located in Katrineholm (170 km south-west of Stockholm),

Fig. 15. Monitored (24 h) average temperature variations (three locations; o

Katrineholm, Sweden in October 2001 to May 2002.

Sweden. A borehole with 83 m active length was used the

heat source. The measurements were taken during a period

of eight months, from 1st of October 2001 to 24th May

2002.

Fig. 15 shows the averaged (24 h) outdoor temperature

(A-ut) variation and the average temperature variation in the

borehole at 25 m (A-BH-25) and 55 m (A-BH-55) depth on

each day. The minimum and maximum average values of

the outdoor temperature were K21 and 17 8C. The

minimum and maximum average values of temperatures

in the borehole at 25 m depth were K0.2 and 5.6 8C and at

55 m depth 0.7 and 5.6 8C.

utdoor, and at two levels in the bore hole, 25 m deep, 55 m deep) in

Fig. 16. Monitored minimum temperature variation (four locations; outdoor, at two levels in bore hole, 25 m deep ground, 55 m deep and brine

inlet temperature to a heat pump) on each day in Katrineholm, Sweden in October 2001 to May 2002.


Fig. 16 shows the minimum outdoor temperature (M-ut);

the minimum temperatures at the two locations of the borehole

(M-BH-25 and M-BH-55) and the minimum brine inlet

temperatures (M-BR-in) to the heat pump on each day. The

range of minimum and maximum values of these outdoor

temperatures were K23 to 12 8C, borehole temperature at

25 m depth,K2 to 5.2 8C, borehole temperature at 55 m depth,

0.5–5 8C and brine inlet temperature to the heat pump,K2.2 to

3.8 8C.

Although the outdoor temperature varied very much

during the period of the eight months, the ground

temperature variation was small. Fig. 16 shows that most

of the time the minimum ground temperature was higher

than the minimum outdoor temperature. These conditions

are favourable for stable operation and high coefficients of

performances of the heat pump.

The minimum brine inlet temperature range from K2.2

to 3.8 8C suggests a best refrigerant charge to the proposed

propane heat pump. Fig. 14 shows that a charge of about

200 g is the best choice for the heat pump providing a COP

between 3.5 and 4.

7. Conclusions

The following conclusions are drawn from the tested

laboratory heat pump and the temperature measurements

from the field heat pump

1.
Use of mini-channel heat exchangers reduces consider-
ably the refrigerant charge in heat pumps and refriger-

ation systems.

2.
The heat transfer coefficients of mini-channel heat
exchangers are high, indicating that the reduction of

charge can be reached without loss of COP.

3.
Careful selection of compressor lubrication oil would
further decrease the refrigerant charge of the heat pump.

4.
Reduction of the tube lengths and exclusion of ball
valves would give an optimum charge for the 5 kW

propane heat pump with ground heat source or other

liquid to liquid type applications of less than 200 g.

Acknowledgements

Financial support from The Swedish National Energy

Administration, aluminium tube supply from Hydro Alu-

minium and compressor supply from Copeland is gratefully

acknowledged. The authors express their gratitude to

Joachim Claesson and Martin Forsen, KTH, for their kind

cooperation.

References

[1] Swedish heat pump association.

[2] Calm J. Ferret Database; 1997.

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[11] ASHRAE handbook. Refrigeration systems and applications,

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propane heat pump with low refrigerant charge: design and laboratory tests

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