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Characteristics of a gravity-assisted heat pipe-based heat exchanger
Heat Recovery Systems & CHP Vol. 11, No. 1, pp. 69-77, 1991 Printed in Great Britain
0890-4332/91 $3.00+ .00 PergamonPress pie
CHARACTERISTICS OF A GRAVITY-ASSISTED PIPE-BASED HEAT EXCHANGER
HEAT
T. WADOWSKI,A . AKBARZADEHand P. JOHNSON Royal Melbourne Institute of Technology, Manufacturing and Process Engineering Department, Melbourne, Australia (Received 29 May 1990)
Abstract--An experimental study of the performance of an air-to-air thermosyphon-based heat exchanger utilizing R-22 as the working fluid has been carried out to investigate its behavior under different operating conditions. A test installation has been developed to model a variety of HVAC real life applications. The results reported in this article describe the influence of various parameters such as: supply and exhaust air stream mass flow rates, stream temperatures and exhaust stream moisture content on the effectiveness of the heat exchangers. Heat exchanger heat flow hysteresis has been recognized. Some optimization criteria are presented.
NOMENCLATURE Cp e E h th m~ n R T w
specificheat of air stream [kJ/kgK] effectivenessof a single row or a single unit consisting of a number of rows effectivenessof a heat exchanger specificenthalpy of air [kJ/kg] mass flow rate [kg/s] the minimum value of supply and exhaust air mass flow rate [kg/s] number of rows or number of units consisting'of a certain number of rows thermal resistance [K/W] air stream temperature [°C] air stream moisture content [kg/kgvRvAm]
Subscripts 1 supply air upstream of the heat exchanger 2 supply air downstream of the heat exchanger 3 exhaust air upstream of the heat exchanger 4 exhaust air downstream of the heat exchanger 4m exhaust air downstream of the heat exchanger relating to equation (9) A dry air s supply e exhaust V water vapour
1. I N T R O D U C T I O N W a s t e heat recovery, one o f the m e t h o d s o f energy c o n s e r v a t i o n , c a n be successfully i m p l e m e n t e d w h e n the i n v e s t m e n t cost o f a d d i t i o n a l e q u i p m e n t required is acceptably low. T h e r m o s y p h o n - b a s e d heat exchangers are very simple products. T h e y m a y be used to transfer heat between two gas streams. F e a t u r e s include n o c r o s s - c o n t a m i n a t i o n between streams, n o m o v i n g parts, c o m p a c t n e s s a n d n o need for a n y external power supply. These features e n c o u r a g e detailed study o f the p r o d u c t ' s p e r f o r m a n c e . T h e schematic o f such a heat recovery system with the c o r r e s p o n d i n g flow a r r a n g e m e n t is s h o w n in Fig. 1.
The thermal performance map of the heat exchanger is created by analysis of its effectiveness as a function of various operating conditions. The effectiveness (E) of a heat exchanger is the ratio of the heat exchanger's rate of heat transferred to the maximum possible rate of heat transfer between the air streams. This can be expressed as follows: E = rhs(h, - h2) #/rain(hi -- h3 ) " 69
(!)
70
T. WADOWSKIet al.
HEAT EXCHANGER /-
J
SUPPLY A[R
rhs
h2
-= h 1
h3
EXHAU$1AIR .~
/%
Fig. 1. Schematic of the flow arrangement and corresponding specific enthalpies.
The specific enthalpy of moist air can be written as: (2)
h = hA ÷ why.
If the moisture contents of both supply and exhaust streams are equal and there is no condensation then the effect of heat transfer is the rise of sensible heat of the supply air stream and therefore equation (1) can be written in simpler form:
,.s(T,- 72) E = rhm~.(T, _ T3)"
(3)
A test installation has been developed to allow for monitoring of the heat exchanger performance under various operating conditions. Tests results presented in the following sections show the influence of various parameters on the effectiveness of different prototypes. The effectiveness hysteresis observed by Stauder and McDonald [1] has been recognized. A proper understanding of the phenomenon is especially important when applying the heat exchangers to air conditioning systems. Some of the work has been carried out to analyze the influence of exhaust stream moisture content on heat exchanger performance. The results presented in this article will serve as the means by which the simulation program developed to model heat exchanger heat transfer will be validated.
Bank of tubes
•$8
~.L O~[~-I j Jl ~
J" 0000 00I
--~
c'~f
Col d st r earn
Condensl sectio
~
~
11/ 0
Cas
Evaporator / section
u~ 0
UUUUUU 30S Fig. 2. Schematic of the developed thermosyphon-based heat exchanger.
Hot stream Fig. 3. Geometry and dimensions of the three-row heat exchanger.
A heat pipe-based heat exchanger
71
2. M E T H O D OF C O N S T R U C T I O N OF THE HEAT E X C H A N G E R A thermosyphon-based heat exchanger is shown in Fig. 2. It consists of a number of finned tubes installed in staggered arrangement in a casing which allows for a flow of two adjacent gases without any cross-contamination. Each tube (thermosyphon) is a separate container initially evacuated and then partially filled with refrigerant. It becomes a thermal superconductor when its lower part is exposed to higher temperature than the upper one. In operation the working fluid evaporates increasing the saturation pressure and temperature in the container. The temperature differential in the upper section causes condensation on the inside wall of the tube. Liquid returns to the evaporator section by gravity. The overall effect is rapid heat transfer, which depends closely on the intensity of the evaporation process. Thermal resistance between the outer tube surface and the gas stream is lowered by extending the contact surface area with rectangular fins. The overall dimensions of the tested prototypes and the tube spacings in the direction of the flow and normal to the flow are shown in Fig. 3. The thermosyphons were made of copper tubes (16.4 mm OD, 15.5 mm ID). Each tube was charged with 50 g of R-22 refrigerant. Corrugated fins were made of aluminum, 0.16 mm thick. Tests were run on units equipped with two different fin spacings: 472 fins/m and 315 fins/re. Overall depth of the heat exchanger varied from 99.6 to 397.2 ram, equivalent to 3 to 12 thermosyphon rows. Each row consisted of eight tubes. 3. TEST FACILITIES The test installation is shown schematically in Fig. 4. Some of its characteristics are listed below: --mass flow rates 0.06 to 0.28 kg/s with an option of varying supply to exhaust stream velocity ratio; --exhaust stream temperature up to 70°C; ---exhaust stream moisture content such that the saturation limit can be reached at temperatures below 70°C. The condenser section of the heat exchanger is located at the suction side of a fan which is driven by a variable speed motor. An air bypass is located downstream of the fan. By opening the bypass damper, exhaust air flow rate can be altered to achieve the desired exhaust to supply stream flow
HEAT EXCHANGER
¥ AIR IN i,
°°°n
1 l
L
i
I
UUUU AIR TREATMENTST.
* ~ D A M PER
BYPASS MEASURINGINsTRuMENTS: dry bulb therm, ~L -wet bu|btherm.
y
- E~Fvetocity
(~ - st.,,c pres.ure
Fig. 4. The constructed rig to determine characteristics of manufactured heat exchangers.
FAN
72
T. WADOWSKI et al.
rate ratio. The temperature and moisture content of the exhaust stream are adjusted by means of a humidifier and electrical heater installed upstream of the evaporator section of the heat exchanger. Performance of the unit is monitored by four measurement stations equipped with wet bulb, dry bulb, and static pressure (where required) measuring facilities. Air mass flow rate is measured at two points. A shaped air inlet assists uniformity of face velocity in the duct entrance. The difference between ambient pressure and static pressure just downstream of the inlet serves as an air flow rate indicator. Hot stream air mass flow rate is measured with a pitot-static tube. Copper-constantan thermocouples are used to measure air temperatures. The data are handled by means of a data acquisition system incorporating a DT100 data logger managed by a specially developed computer program run from an IBM PC. The program automatically processes the data to provide information such as effectiveness and the heat transfer rate. The bypass is connected to the outlet duct from which the air is exhausted to the atmosphere. Both thermocouples and mass flow rate measuring devices were calibrated against a reference thermometer and a standard Venturi tube. As a result the temperature measurements tolerance is 0.4°C. Mass flow rate measurement accuracy is 6% of the smallest mass flow rate considered and decreases with higher values. Face velocity profile has been determined by means of nine pitot-static tubes installed in the duct. The velocity deviations along the cross-sectional area of the duct were smaller than 7% of the mean value. A grid of nine thermocouples ser,Jed to analyze temperature distribution in the duct. The temperature deviations in extreme cases did not exceed 2'~C. 4. RESULTS Air stream properties such as mass flow rates, mass flow rate ratio, stream temperatures, moisture content and heat exchanger geometry can influence performance of the unit. In this study the work has been concentrated on recreation of conditions supplied by HVAC systems and therefore the results are presented for stream temperatures not exceeding 70°C and face velocity not higher than 2.5 m/s.
4.1 The effect of mass flow rate The effect of change of mass flow rate of both exhaust and supply stream is shown in Figs 5 and 6. The temperatures of exhaust and supply air streams have been kept constant throughout 100 go 80 70 (n
60
z 50
bJ
40 30 2O I0 0
I
0.04-
I
0.08
I
!
!
0.12
MASS R.OW ~
I
I
0,16
I
0.2
I
I
I
0.24-
[ke/,]
Fig. 5. Effectiveness as a function of air streams mass flow rates shown for 472 fins/m spacing.
0.28
A heat pipe-bascdheat exchanger
73
100 90 80 70r--1 t_.l
6050 40 30 20 10 0
I
0.04.
I
I
0.08
I
I
O. 12
I
I
0.16
I
0.2
I
I
I
0.24
0.28
r o s s n_oR ~ATE [ k Q / , l Fig. 6. Effectiveness as a function of air streams mass flow rates shown for 315 fins/m spacing.
the experiment at 50 and 20°C, respectively. Moisture content was the same for both air streams and changed with meteorological conditions. The mass flow rates have been varied in the range of 0.06-0.28 kg/s. The two charts represent units equipped with different fin spacing. Each curve represents different heat exchanger overall depth, characterized by the number of thermosyphon rows. The effectiveness value drops with the rise of air mass flow rates. The simplest way to find the reason for such behavior is an analysis of the simplest heat exchanger model consisting of one thermal resistance (R). A correlation describing energy balance of the heat exchanger exposed to two streams of air (see Fig. 1) at different temperatures can be developed on the basis of such an assumption: (T3 -
T2)/R = Cprhs(T2- I"i).
(4)
If ms = the and since the moisture content of both streams is the same then equation (3) becomes: E = (7"2- Tt)
( T 3 - T,)"
(5)
After substitution to (4) the effectiveness can be represented by:
E
= 1/(CprhsR + l)
(6)
and the drop of effectiveness with rise of mass flow rate is obviously the case. (It can be seen that when rh, decreases, the effectiveness (E) increases and approaches unity. On the other hand when rhs approaches infinity E becomes zero.) 4.2 The effect o f heat exchanger depth and the fin spacing Effectiveness value increases with increasing depth of the heat exchanger. The effect of change of heat exchanger depth (characterized by number of thermosyphon rows) is also shown in Figs 5 and 6. Points of constant mass flow rate corresponding to particular depths of the heat exchanger are the means by which this influence may be analyzed. It is also worth noticing that the effectiveness value can be approximated by the equation presented in [6]: E = n e / ( l + ( n - l)e). HRS II/I--F
(7)
74
T. WADOWSKI et al.
The above equation is commonly used when analyzing performance of heat exchangers consisting of a number of combined units. There is a good agreement between the experimental data and the results obtained using equation (7). When comparing the results presented in 'the charts, it can be seen that the slight change of effectiveness value can be accounted for by the change of fin spacing. The effectiveness ratio is almost a constant value equal to 0.86 where the fin surface area ratio, which is equal to fin spacing ratio, is 0.67. When altering fin spacing the thermal resistance of finned area is affected by the change of surface area and also the convection heat transfer coefficient. 4.3 The effect of mass flow rate ratio The changes of mass flow rate ratio would be expected to have a substantial effect on the effectiveness value. Equation (1) indicates its significance. The effect of unequal mass flow rate ratio has been investigated. The results presented in Fig. 7 match with ones already available describing similar type of heat exchangers performance [4, 7]. The test was carried out on a six-row heat exchanger with 472 fins/m fin spacing. Supply and exhaust stream temperatures were maintained at 22 and 42°C and the moisture contents of streams were kept constant throughout the experiment. The supply stream has been kept at 0.28 kg/s mass flow rate and the exhaust stream flow rate was changed to obtain ratios ranging from 1 to 2.5. From tests results it is apparent that with rise of mass flow rate ratio the effectiveness rises. 4.4. The effect of exhaust stream moisture content There are heat exchanger applications featuring unequal supply and exhaust stream moisture content. Dehumidifiers, for instance, exhaust humid air and in fact only sensible heat may be recovered when preheating the supply air stream. Furthermore, in the case of unequal stream moisture contents, equation (1) describing the effectiveness must be altered. The maximum possible rate of heat transfer is equal to: ~max = /~min (h3 - - h4m)
(8)
where h4u represents the enthalpy of moist air at its lowest possible temperature (equivalent to
100 go 80 70 60 50 40 30 20 10 0
i I
I
1.2
!
!
1.4.
!
I
I
1.6
I
1.8
I
l
2
!
!
!
2.2
FLOW RATE[ RKIIO
Fig. 7. Effectiveness as a function of mass flow rate ratio of a six-row heat exchanger.
!
2.4-
A heat pipe-based heat exchanger
75
100 g0 80 70 60 id
SENSIBLE
50
--
~
n
[]
i
.,CONDENSATION REGION 0
I~ A
4O id 30 -~ 20 10
I !
0
I
10
I
I
|
30
f
50
M ~ S ~ R E CONTENT
70
ra~d
Fig. 8. Effectiveness as a function of exhaust stream moisture content for a six-row heat exchanger.
supply stream temperature) and moisture content corresponding to the actual exhaust stream humidity. The effectiveness becomes
E=
rhs(h2-hl) thmin(h3 - h,m)"
(9)
The above equation is especially valuable when the exhaust stream at temperature T~ reaches saturation limit or is beyond it and part of the vapour condenses. Then the enthalpy difference in the denominator can be based on two different moisture contents. No condensation can take place in the condenser section of the heat exchanger since the supply stream is being heated up. The numerator stays the same. Equation (9) is not as practical as the sensible effectiveness relationship (3). It is suggested to use equation (3) for comparative study of the heat exchanger performance since if the mass flow rate is constant then the denominator of equation (3) stays the same. This way any change in heat exchanger heat transfer rate can be easily detected. The influence of the exhaust stream moisture content on the heat exchanger heat transfer and its effectiveness has been investigated using a six-row heat exchanger having 472 fins/m fin spacing. Both exhaust and supply air stream mass flow rates were kept constant and equal to 0.28 kg/s. The exhaust stream temperature was maintained at 50°C and the supply stream at 25°C. The supply stream moisture content was equal to 10 g/kg and the exhaust stream moisture content was changed between 10 and 70 g/kg. The results of the test are shown in Fig. 8. There are two curves representing the sensible effectiveness as calculated from equation (3), and total effectiveness as a function of exhaust stream moisture content as calculated from equation (9). It can be seen that the sensible effectiveness and therefore the actual heat transfer rate is almost a constant value independent of the exhaust air moisture content. At very high exhaust stream moisture content the sensible effectiveness slightly increases. Heat transfer between air and the fins in heat exchanger evaporator probably improves because an increase of humidity ratio means an increase of air stream density. Water condensation taking place on the surface of the fins may also account for higher heat transfer since water droplets increase contact surface area between air and the heat sink. The total effectiveness is a constant value until the moisture content of the exhaust stream reaches 20 g/kg. This is sufficient for the air to reach the saturation limit at temperature T~. Further increase of moisture content causes a rapid drop of total effectiveness value. The actual condensation of
T . WADOWSKI et al.
76
64
62 60 FALLING TEMP. OIFF
58 56 54 52 w
Z
50
w
48
0 w h tl_ W
46 44 42 40 38
!
36 0
20
40
60
TEMPERATURE DIFFERENCE [deg. C] F i g . 9.
The effectiveness a s a function of temperature difference for a six-row heat exchanger (equal supply and exhaust air stream mass flow rates).
water began when exhaust stream moisture content was equal to 40 g/kg. That did not change the behavior of total effectiveness value. One way to improve the total effectiveness and recover heat present in the form of latent heat is to raise the supply to exhaust stream mass flow rate ratio. 4.5. The effect of stream temperature difference
It has been found that the effectiveness value depends on the thermal history of the heat exchanger. The hysteresis shown in Fig. 9, obtained for a six-row heat exchanger at 0.17 kg/s mass flow rate illustrates its importance. Uneven performance of the heat exchanger can be explained by the fact that the working fluid present in the thermosyphon must be superheated to initiate the boiling process [2]. When that happens a rapid heat transfer begins. A sharper "jump" of effectiveness does not occur since different rows of tubes in the exchanger are exposed to different temperature differentials. It can be seen that full operation of the heat exchanger is reached when the difference between supply and exhaust stream temperature is 30°C. Once started, the boiling process is sustained for even a very small temperature difference. In fact, even at as low as 5°C temperature differential the system keeps a constant effectiveness of 54% indefinitely. It is also worth noticing that there is almost no influence of the temperature differential on the effectiveness when the system has initially been preheated above 30°C and then kept above the limit of 5°C. 5. DISCUSSION The hysteresis effect should always be taken into account when the heat exchanger is used in applications where small temperature differences are expected. The latent heat of water present in moist exhaust air streams from dehumidifiers may be recovered when air stream flow rates much larger than exhaust stream flow rates are passed through the condenser section of the heat exchanger. Dry air has got much lower thermal potential to absorb heat than exhausted moist air to reject it. When optimizing the heat exchanger for particular applications, factors such as investment and running cost must be compared with the savings from heat recovery. It is possible to increase effectiveness of the heat,exchanger by decreasing face velocity of either one stream or both air streams. This can be done by increasing the cross-sectional area of the heat exchanger and the
A heat pipe-based heat exchanger
77
ducting. Such alteration will raise the investment cost. Furthermore, it is possible to increase the overall depth of the heat exchanger. Additional pressure losses account for increase in the running cost which must further increase investment cost. 6. C O N C L U S I O N S
A minimum temperature difference between the two air streams is required to initiate operation of the heat exchanger. When full operating power is reached the effectiveness is independent of temperature difference of the two air streams. Effectiveness of the heat exchanger installed in the stream of moist exhausted air can be estimated on the basis of calculation of an ideal effectiveness for a particular condition. It has been found that the performance of the heat exchanger does not improve with change of air stream density and with condensation. A full performance map is available for units equipped with fin spacing 472 and 315 fins/m. REFERENCES 1. F. A. Stauder and T. W. McDonald, Experimental Study of a Two Phase Thermosyphon Loop Heat Exchanger, A S H R A E Trans. 92 (2A), 486-487 (1986). 2. R. Cole, Boiling Nucleation, Advances of Heat Transfer, Academic Press, 10, (1974). 3. D. Ge, T. W. McDonnald and G. D. Mathur, Hysteresis in Two Phase Thermosyphon Loop Heat Exchangers, ASHRAE Trans. 93, 275-281 (1987). 4. Y. Lee and A. Bedrossian, The Characteristics of Heat Exchangers Using Heat Pipes or Thermosyphons. Int. J. Heat Mass Trans. 21, 221-229, Pergamon Press (1978). 5. E. Azad and F. Geoola, A Design Procedure for Gravity Assisted Heat Pipe Heat Exchanger, J. Heat Recovery Systems 4, 101-111 (1984). 6. W. M. Keys and A. L. London, Compact Heat Exchangers, McGraw-Hill (1984). 7. J. J. Bosh and G. J. Gudac, Effectiveness and Pressure Drop Characteristic of Various Types of Air-to-air Energy Recovery Systems. A S H R A E Trans. 87, 199-210 (1981). 8. Method of Testing Air-to-air Heat Exchangers, ASHRAE Standard, 84-78.
0890-4332/91 $3.00+ .00 PergamonPress pie
CHARACTERISTICS OF A GRAVITY-ASSISTED PIPE-BASED HEAT EXCHANGER
HEAT
T. WADOWSKI,A . AKBARZADEHand P. JOHNSON Royal Melbourne Institute of Technology, Manufacturing and Process Engineering Department, Melbourne, Australia (Received 29 May 1990)
Abstract--An experimental study of the performance of an air-to-air thermosyphon-based heat exchanger utilizing R-22 as the working fluid has been carried out to investigate its behavior under different operating conditions. A test installation has been developed to model a variety of HVAC real life applications. The results reported in this article describe the influence of various parameters such as: supply and exhaust air stream mass flow rates, stream temperatures and exhaust stream moisture content on the effectiveness of the heat exchangers. Heat exchanger heat flow hysteresis has been recognized. Some optimization criteria are presented.
NOMENCLATURE Cp e E h th m~ n R T w
specificheat of air stream [kJ/kgK] effectivenessof a single row or a single unit consisting of a number of rows effectivenessof a heat exchanger specificenthalpy of air [kJ/kg] mass flow rate [kg/s] the minimum value of supply and exhaust air mass flow rate [kg/s] number of rows or number of units consisting'of a certain number of rows thermal resistance [K/W] air stream temperature [°C] air stream moisture content [kg/kgvRvAm]
Subscripts 1 supply air upstream of the heat exchanger 2 supply air downstream of the heat exchanger 3 exhaust air upstream of the heat exchanger 4 exhaust air downstream of the heat exchanger 4m exhaust air downstream of the heat exchanger relating to equation (9) A dry air s supply e exhaust V water vapour
1. I N T R O D U C T I O N W a s t e heat recovery, one o f the m e t h o d s o f energy c o n s e r v a t i o n , c a n be successfully i m p l e m e n t e d w h e n the i n v e s t m e n t cost o f a d d i t i o n a l e q u i p m e n t required is acceptably low. T h e r m o s y p h o n - b a s e d heat exchangers are very simple products. T h e y m a y be used to transfer heat between two gas streams. F e a t u r e s include n o c r o s s - c o n t a m i n a t i o n between streams, n o m o v i n g parts, c o m p a c t n e s s a n d n o need for a n y external power supply. These features e n c o u r a g e detailed study o f the p r o d u c t ' s p e r f o r m a n c e . T h e schematic o f such a heat recovery system with the c o r r e s p o n d i n g flow a r r a n g e m e n t is s h o w n in Fig. 1.
The thermal performance map of the heat exchanger is created by analysis of its effectiveness as a function of various operating conditions. The effectiveness (E) of a heat exchanger is the ratio of the heat exchanger's rate of heat transferred to the maximum possible rate of heat transfer between the air streams. This can be expressed as follows: E = rhs(h, - h2) #/rain(hi -- h3 ) " 69
(!)
70
T. WADOWSKIet al.
HEAT EXCHANGER /-
J
SUPPLY A[R
rhs
h2
-= h 1
h3
EXHAU$1AIR .~
/%
Fig. 1. Schematic of the flow arrangement and corresponding specific enthalpies.
The specific enthalpy of moist air can be written as: (2)
h = hA ÷ why.
If the moisture contents of both supply and exhaust streams are equal and there is no condensation then the effect of heat transfer is the rise of sensible heat of the supply air stream and therefore equation (1) can be written in simpler form:
,.s(T,- 72) E = rhm~.(T, _ T3)"
(3)
A test installation has been developed to allow for monitoring of the heat exchanger performance under various operating conditions. Tests results presented in the following sections show the influence of various parameters on the effectiveness of different prototypes. The effectiveness hysteresis observed by Stauder and McDonald [1] has been recognized. A proper understanding of the phenomenon is especially important when applying the heat exchangers to air conditioning systems. Some of the work has been carried out to analyze the influence of exhaust stream moisture content on heat exchanger performance. The results presented in this article will serve as the means by which the simulation program developed to model heat exchanger heat transfer will be validated.
Bank of tubes
•$8
~.L O~[~-I j Jl ~
J" 0000 00I
--~
c'~f
Col d st r earn
Condensl sectio
~
~
11/ 0
Cas
Evaporator / section
u~ 0
UUUUUU 30S Fig. 2. Schematic of the developed thermosyphon-based heat exchanger.
Hot stream Fig. 3. Geometry and dimensions of the three-row heat exchanger.
A heat pipe-based heat exchanger
71
2. M E T H O D OF C O N S T R U C T I O N OF THE HEAT E X C H A N G E R A thermosyphon-based heat exchanger is shown in Fig. 2. It consists of a number of finned tubes installed in staggered arrangement in a casing which allows for a flow of two adjacent gases without any cross-contamination. Each tube (thermosyphon) is a separate container initially evacuated and then partially filled with refrigerant. It becomes a thermal superconductor when its lower part is exposed to higher temperature than the upper one. In operation the working fluid evaporates increasing the saturation pressure and temperature in the container. The temperature differential in the upper section causes condensation on the inside wall of the tube. Liquid returns to the evaporator section by gravity. The overall effect is rapid heat transfer, which depends closely on the intensity of the evaporation process. Thermal resistance between the outer tube surface and the gas stream is lowered by extending the contact surface area with rectangular fins. The overall dimensions of the tested prototypes and the tube spacings in the direction of the flow and normal to the flow are shown in Fig. 3. The thermosyphons were made of copper tubes (16.4 mm OD, 15.5 mm ID). Each tube was charged with 50 g of R-22 refrigerant. Corrugated fins were made of aluminum, 0.16 mm thick. Tests were run on units equipped with two different fin spacings: 472 fins/m and 315 fins/re. Overall depth of the heat exchanger varied from 99.6 to 397.2 ram, equivalent to 3 to 12 thermosyphon rows. Each row consisted of eight tubes. 3. TEST FACILITIES The test installation is shown schematically in Fig. 4. Some of its characteristics are listed below: --mass flow rates 0.06 to 0.28 kg/s with an option of varying supply to exhaust stream velocity ratio; --exhaust stream temperature up to 70°C; ---exhaust stream moisture content such that the saturation limit can be reached at temperatures below 70°C. The condenser section of the heat exchanger is located at the suction side of a fan which is driven by a variable speed motor. An air bypass is located downstream of the fan. By opening the bypass damper, exhaust air flow rate can be altered to achieve the desired exhaust to supply stream flow
HEAT EXCHANGER
¥ AIR IN i,
°°°n
1 l
L
i
I
UUUU AIR TREATMENTST.
* ~ D A M PER
BYPASS MEASURINGINsTRuMENTS: dry bulb therm, ~L -wet bu|btherm.
y
- E~Fvetocity
(~ - st.,,c pres.ure
Fig. 4. The constructed rig to determine characteristics of manufactured heat exchangers.
FAN
72
T. WADOWSKI et al.
rate ratio. The temperature and moisture content of the exhaust stream are adjusted by means of a humidifier and electrical heater installed upstream of the evaporator section of the heat exchanger. Performance of the unit is monitored by four measurement stations equipped with wet bulb, dry bulb, and static pressure (where required) measuring facilities. Air mass flow rate is measured at two points. A shaped air inlet assists uniformity of face velocity in the duct entrance. The difference between ambient pressure and static pressure just downstream of the inlet serves as an air flow rate indicator. Hot stream air mass flow rate is measured with a pitot-static tube. Copper-constantan thermocouples are used to measure air temperatures. The data are handled by means of a data acquisition system incorporating a DT100 data logger managed by a specially developed computer program run from an IBM PC. The program automatically processes the data to provide information such as effectiveness and the heat transfer rate. The bypass is connected to the outlet duct from which the air is exhausted to the atmosphere. Both thermocouples and mass flow rate measuring devices were calibrated against a reference thermometer and a standard Venturi tube. As a result the temperature measurements tolerance is 0.4°C. Mass flow rate measurement accuracy is 6% of the smallest mass flow rate considered and decreases with higher values. Face velocity profile has been determined by means of nine pitot-static tubes installed in the duct. The velocity deviations along the cross-sectional area of the duct were smaller than 7% of the mean value. A grid of nine thermocouples ser,Jed to analyze temperature distribution in the duct. The temperature deviations in extreme cases did not exceed 2'~C. 4. RESULTS Air stream properties such as mass flow rates, mass flow rate ratio, stream temperatures, moisture content and heat exchanger geometry can influence performance of the unit. In this study the work has been concentrated on recreation of conditions supplied by HVAC systems and therefore the results are presented for stream temperatures not exceeding 70°C and face velocity not higher than 2.5 m/s.
4.1 The effect of mass flow rate The effect of change of mass flow rate of both exhaust and supply stream is shown in Figs 5 and 6. The temperatures of exhaust and supply air streams have been kept constant throughout 100 go 80 70 (n
60
z 50
bJ
40 30 2O I0 0
I
0.04-
I
0.08
I
!
!
0.12
MASS R.OW ~
I
I
0,16
I
0.2
I
I
I
0.24-
[ke/,]
Fig. 5. Effectiveness as a function of air streams mass flow rates shown for 472 fins/m spacing.
0.28
A heat pipe-bascdheat exchanger
73
100 90 80 70r--1 t_.l
6050 40 30 20 10 0
I
0.04.
I
I
0.08
I
I
O. 12
I
I
0.16
I
0.2
I
I
I
0.24
0.28
r o s s n_oR ~ATE [ k Q / , l Fig. 6. Effectiveness as a function of air streams mass flow rates shown for 315 fins/m spacing.
the experiment at 50 and 20°C, respectively. Moisture content was the same for both air streams and changed with meteorological conditions. The mass flow rates have been varied in the range of 0.06-0.28 kg/s. The two charts represent units equipped with different fin spacing. Each curve represents different heat exchanger overall depth, characterized by the number of thermosyphon rows. The effectiveness value drops with the rise of air mass flow rates. The simplest way to find the reason for such behavior is an analysis of the simplest heat exchanger model consisting of one thermal resistance (R). A correlation describing energy balance of the heat exchanger exposed to two streams of air (see Fig. 1) at different temperatures can be developed on the basis of such an assumption: (T3 -
T2)/R = Cprhs(T2- I"i).
(4)
If ms = the and since the moisture content of both streams is the same then equation (3) becomes: E = (7"2- Tt)
( T 3 - T,)"
(5)
After substitution to (4) the effectiveness can be represented by:
E
= 1/(CprhsR + l)
(6)
and the drop of effectiveness with rise of mass flow rate is obviously the case. (It can be seen that when rh, decreases, the effectiveness (E) increases and approaches unity. On the other hand when rhs approaches infinity E becomes zero.) 4.2 The effect o f heat exchanger depth and the fin spacing Effectiveness value increases with increasing depth of the heat exchanger. The effect of change of heat exchanger depth (characterized by number of thermosyphon rows) is also shown in Figs 5 and 6. Points of constant mass flow rate corresponding to particular depths of the heat exchanger are the means by which this influence may be analyzed. It is also worth noticing that the effectiveness value can be approximated by the equation presented in [6]: E = n e / ( l + ( n - l)e). HRS II/I--F
(7)
74
T. WADOWSKI et al.
The above equation is commonly used when analyzing performance of heat exchangers consisting of a number of combined units. There is a good agreement between the experimental data and the results obtained using equation (7). When comparing the results presented in 'the charts, it can be seen that the slight change of effectiveness value can be accounted for by the change of fin spacing. The effectiveness ratio is almost a constant value equal to 0.86 where the fin surface area ratio, which is equal to fin spacing ratio, is 0.67. When altering fin spacing the thermal resistance of finned area is affected by the change of surface area and also the convection heat transfer coefficient. 4.3 The effect of mass flow rate ratio The changes of mass flow rate ratio would be expected to have a substantial effect on the effectiveness value. Equation (1) indicates its significance. The effect of unequal mass flow rate ratio has been investigated. The results presented in Fig. 7 match with ones already available describing similar type of heat exchangers performance [4, 7]. The test was carried out on a six-row heat exchanger with 472 fins/m fin spacing. Supply and exhaust stream temperatures were maintained at 22 and 42°C and the moisture contents of streams were kept constant throughout the experiment. The supply stream has been kept at 0.28 kg/s mass flow rate and the exhaust stream flow rate was changed to obtain ratios ranging from 1 to 2.5. From tests results it is apparent that with rise of mass flow rate ratio the effectiveness rises. 4.4. The effect of exhaust stream moisture content There are heat exchanger applications featuring unequal supply and exhaust stream moisture content. Dehumidifiers, for instance, exhaust humid air and in fact only sensible heat may be recovered when preheating the supply air stream. Furthermore, in the case of unequal stream moisture contents, equation (1) describing the effectiveness must be altered. The maximum possible rate of heat transfer is equal to: ~max = /~min (h3 - - h4m)
(8)
where h4u represents the enthalpy of moist air at its lowest possible temperature (equivalent to
100 go 80 70 60 50 40 30 20 10 0
i I
I
1.2
!
!
1.4.
!
I
I
1.6
I
1.8
I
l
2
!
!
!
2.2
FLOW RATE[ RKIIO
Fig. 7. Effectiveness as a function of mass flow rate ratio of a six-row heat exchanger.
!
2.4-
A heat pipe-based heat exchanger
75
100 g0 80 70 60 id
SENSIBLE
50
--
~
n
[]
i
.,CONDENSATION REGION 0
I~ A
4O id 30 -~ 20 10
I !
0
I
10
I
I
|
30
f
50
M ~ S ~ R E CONTENT
70
ra~d
Fig. 8. Effectiveness as a function of exhaust stream moisture content for a six-row heat exchanger.
supply stream temperature) and moisture content corresponding to the actual exhaust stream humidity. The effectiveness becomes
E=
rhs(h2-hl) thmin(h3 - h,m)"
(9)
The above equation is especially valuable when the exhaust stream at temperature T~ reaches saturation limit or is beyond it and part of the vapour condenses. Then the enthalpy difference in the denominator can be based on two different moisture contents. No condensation can take place in the condenser section of the heat exchanger since the supply stream is being heated up. The numerator stays the same. Equation (9) is not as practical as the sensible effectiveness relationship (3). It is suggested to use equation (3) for comparative study of the heat exchanger performance since if the mass flow rate is constant then the denominator of equation (3) stays the same. This way any change in heat exchanger heat transfer rate can be easily detected. The influence of the exhaust stream moisture content on the heat exchanger heat transfer and its effectiveness has been investigated using a six-row heat exchanger having 472 fins/m fin spacing. Both exhaust and supply air stream mass flow rates were kept constant and equal to 0.28 kg/s. The exhaust stream temperature was maintained at 50°C and the supply stream at 25°C. The supply stream moisture content was equal to 10 g/kg and the exhaust stream moisture content was changed between 10 and 70 g/kg. The results of the test are shown in Fig. 8. There are two curves representing the sensible effectiveness as calculated from equation (3), and total effectiveness as a function of exhaust stream moisture content as calculated from equation (9). It can be seen that the sensible effectiveness and therefore the actual heat transfer rate is almost a constant value independent of the exhaust air moisture content. At very high exhaust stream moisture content the sensible effectiveness slightly increases. Heat transfer between air and the fins in heat exchanger evaporator probably improves because an increase of humidity ratio means an increase of air stream density. Water condensation taking place on the surface of the fins may also account for higher heat transfer since water droplets increase contact surface area between air and the heat sink. The total effectiveness is a constant value until the moisture content of the exhaust stream reaches 20 g/kg. This is sufficient for the air to reach the saturation limit at temperature T~. Further increase of moisture content causes a rapid drop of total effectiveness value. The actual condensation of
T . WADOWSKI et al.
76
64
62 60 FALLING TEMP. OIFF
58 56 54 52 w
Z
50
w
48
0 w h tl_ W
46 44 42 40 38
!
36 0
20
40
60
TEMPERATURE DIFFERENCE [deg. C] F i g . 9.
The effectiveness a s a function of temperature difference for a six-row heat exchanger (equal supply and exhaust air stream mass flow rates).
water began when exhaust stream moisture content was equal to 40 g/kg. That did not change the behavior of total effectiveness value. One way to improve the total effectiveness and recover heat present in the form of latent heat is to raise the supply to exhaust stream mass flow rate ratio. 4.5. The effect of stream temperature difference
It has been found that the effectiveness value depends on the thermal history of the heat exchanger. The hysteresis shown in Fig. 9, obtained for a six-row heat exchanger at 0.17 kg/s mass flow rate illustrates its importance. Uneven performance of the heat exchanger can be explained by the fact that the working fluid present in the thermosyphon must be superheated to initiate the boiling process [2]. When that happens a rapid heat transfer begins. A sharper "jump" of effectiveness does not occur since different rows of tubes in the exchanger are exposed to different temperature differentials. It can be seen that full operation of the heat exchanger is reached when the difference between supply and exhaust stream temperature is 30°C. Once started, the boiling process is sustained for even a very small temperature difference. In fact, even at as low as 5°C temperature differential the system keeps a constant effectiveness of 54% indefinitely. It is also worth noticing that there is almost no influence of the temperature differential on the effectiveness when the system has initially been preheated above 30°C and then kept above the limit of 5°C. 5. DISCUSSION The hysteresis effect should always be taken into account when the heat exchanger is used in applications where small temperature differences are expected. The latent heat of water present in moist exhaust air streams from dehumidifiers may be recovered when air stream flow rates much larger than exhaust stream flow rates are passed through the condenser section of the heat exchanger. Dry air has got much lower thermal potential to absorb heat than exhausted moist air to reject it. When optimizing the heat exchanger for particular applications, factors such as investment and running cost must be compared with the savings from heat recovery. It is possible to increase effectiveness of the heat,exchanger by decreasing face velocity of either one stream or both air streams. This can be done by increasing the cross-sectional area of the heat exchanger and the
A heat pipe-based heat exchanger
77
ducting. Such alteration will raise the investment cost. Furthermore, it is possible to increase the overall depth of the heat exchanger. Additional pressure losses account for increase in the running cost which must further increase investment cost. 6. C O N C L U S I O N S
A minimum temperature difference between the two air streams is required to initiate operation of the heat exchanger. When full operating power is reached the effectiveness is independent of temperature difference of the two air streams. Effectiveness of the heat exchanger installed in the stream of moist exhausted air can be estimated on the basis of calculation of an ideal effectiveness for a particular condition. It has been found that the performance of the heat exchanger does not improve with change of air stream density and with condensation. A full performance map is available for units equipped with fin spacing 472 and 315 fins/m. REFERENCES 1. F. A. Stauder and T. W. McDonald, Experimental Study of a Two Phase Thermosyphon Loop Heat Exchanger, A S H R A E Trans. 92 (2A), 486-487 (1986). 2. R. Cole, Boiling Nucleation, Advances of Heat Transfer, Academic Press, 10, (1974). 3. D. Ge, T. W. McDonnald and G. D. Mathur, Hysteresis in Two Phase Thermosyphon Loop Heat Exchangers, ASHRAE Trans. 93, 275-281 (1987). 4. Y. Lee and A. Bedrossian, The Characteristics of Heat Exchangers Using Heat Pipes or Thermosyphons. Int. J. Heat Mass Trans. 21, 221-229, Pergamon Press (1978). 5. E. Azad and F. Geoola, A Design Procedure for Gravity Assisted Heat Pipe Heat Exchanger, J. Heat Recovery Systems 4, 101-111 (1984). 6. W. M. Keys and A. L. London, Compact Heat Exchangers, McGraw-Hill (1984). 7. J. J. Bosh and G. J. Gudac, Effectiveness and Pressure Drop Characteristic of Various Types of Air-to-air Energy Recovery Systems. A S H R A E Trans. 87, 199-210 (1981). 8. Method of Testing Air-to-air Heat Exchangers, ASHRAE Standard, 84-78.