Compact heat and mass exchangers of the plate fin type in thermal sorption systems

COMPACT HEAT AND MASS EXCHANGERS OF THE

PLATE FIN TYPE IN THERMAL SORPTION SYSTEMS

Application in an absorption heat pump

with the working pair CH30H-LiBr/ZnBr2

COMPACT HEAT AND MASS EXCHANGERS OF THE

PLATE FIN TYPE IN THERMAL SORPTION SYSTEMS

Application in an absorption heat pump

with the working pair CH30H-LiBr/ZnBr2

CIP-DATA KONINKLIJKE BIBLIOTHEEK, DEN HAAG Becker, Harry

Compact heat and mass exchangers of the plate fin type in thermal sorption systems: appiication in an absorption heat pump with the working pair CH30H-LiBr/ZnBr2

Harry Becker. - Delft: Delft University of Technology, Mechanical Engineering Department Thesis Delft. - With ref. - With summary in Dutch

ISBN 90-370-0022-3

SISO 653.3 UDC 621.577(043.3)

Subject heading: absorption heat pumps.

Copyright © 1989, Faculty of Mechanical Engineering and Marine Engineering Delft University of Technology

All rights reserved.

This report, or parts thereof, may not be reproduced in any form without permission of the publisher.

COMPACT HEAT AND MASS EXCHANGERS OF THE

PLATE FIN TYPE IN THERMAL SORPTION SYSTEMS

Application in an absorption heat pump

with the working pair CH30H-LiBr/ZnBr2

PROEFSCHRIFT

Ter verkrijging van de graad van doctor aan de Technische Universiteit Delft, op gezag van de Rector Magnificus, Prof.drs. P.A. Schenck

in het openbaar te verdedigen ten overstaan van een commissie aangewezen door het College van Dekanen

op dinsdag 7 februari 1989 om 14.00 uur.

door

HARRY BECKER

geboren te Amersfoort Werktuigkundig ingenieur

Dit proefschrift is goedgekeurd door de promotor

CONTENTS SUMMARY

SAMENVATTING

CHAPTER 1. THE ABSORPTION HEAT PUMP 1.1 Introduction

1.2 Working principle

1.3 The log P - l/T diagram

1.4 The heat ratio, the coëfficiënt of performance and the circulation ratio

1.5 Research work on absorption heat pumps (AHP) 1.5.1 Working pairs for an AHP

1.5.2 System configurations 1.5.3 Heat and mass transfer 1.6 This research work

1.6.1 Goals 1.6.2 Tools 1.6.3 Set up

CHAPTER 2. COMPACT HEAT AND MASS EXCHANGERS

page 1 1 3 4 6 6 8 10 11 12 12

2.4

1 Definition

2 Compact and enhanced transfer surface 3 Application 2.3.1 Field of application 2.3.2 Corrugations 2.3.3 Construction materials Characterization 2.4.1 Identification 2.4.2 Analytical solutions 2.4.3 Experimental results 2.4.3.1 Introduction

2.4.3.2 The work of Kays and London 2.4.3.3 Rectangular offset

strip fin surfaces Discussion

Application in sorption systems

CHAPTER 3. THE ABSORPTION HEAT PUMP (AHP) TEST PLANT

3.1 Introduction

3.2 Choice of the working pair

3.3 Type of compact heat and/or mass transfer surface 3.4 Overview of the test plant

3.5 The components

3.5.1 The absorber

3.5.2 The condenser and evaporator

3.5.3 The mixture-mixture heat exchanger 3.5.4 The generator 13 13 15 15 16 16 17 18 18 18 21 22 25 26 26 27 29 30 31 31 31

3.6

The heating and cooling circuits

3.6.1 The cooling system of the absorber and condenser

6.2 The heating system of the evaporator 6.3 The heating system of the generator 6.4 General

3.7

3 3 3 Control 3 3 3 3 7.1 Mass flows 7.2 Temperatures 7.3 Weight fraction 7.4 Pressures 3.8 Measurement 3.8.1 Mass flows 3.8.2 3.8.3 3.8.4 3.8.5 3.8.6 Mixture density Temperatures Pressures Accuracy of measurements Data registration

3.9 Construction materials and corrosion

CHAPTER 4. THE WORKING PAIR : SELECTION AND DATA 4.1 Introduction

4.2 Experiments 4.3 Simulation

4.4 Literature survey on working pairs 4.5 Selection of a new working pair 4.6 Consequences for the test plant

4.7 Consequences for the simulation program

4.8 Comparison R123a - DTG with CH30H - LiBr / ZnBr2

4.9 Conclusions 32 32 32 32 33 33 33 34 34 34 34 34 35 36 36 39 39 39 39 40 41 41 41 43 CHAPTER 5. COMPUTER SIMULATION MODEL AND PROGRAM

5.1 Introduction 45 5.2 Starting point 45 5.3 Development of a new simulation model and program 46

5.4 Final simulation program

5.4.1 Structure 47 5.4.2 Heat and/or mass transfer correlations 48

5.4.3 Computer subroutines 50 5.5 A simplified simulation program 51 5.6 Application of the simulation program: Some examples 52

5.7 Conclusions 55

CHAPTER 6. RESULTS OF THE EXPERIMENTS

6.1 Introduction 57 6.2 Experimental results of the AHP and its components

6.2.2

The

components 6.2.2.1 The 6.2.2.2 The 6.2.2.3 The 6.2.2.4 The 6.2.2.5 The generator condenser evaporator absorber

mixt.-mixt. heat exchanger 60 61 63 65

70

6.3 Conclusions from the experiments 71

CHAPTER 7. INTERPRETATION OF AND DISCUSSION ON THE EXPERIMENTAL RESULTS

7.1 Introduction 73 7.2 A qualitative interpretation 73

7.3 A quantitative interpretation

7.3.1 Introduction 74 7.3.2 The mixture-mixture heat exchanger 75

7.3.3 The absorber 77 7.3.4 Discussion and comparison 79

7.4 Detailed absorber simulation model

7.4.1 Introduction 81 7.4.2 Former researchers 82

7.4.3 Results from this research work 83 7.4.4 Discussion on correction factors 84 7.5 Comparison SAT absorber/generator and the CHME absorber

7.5.1 Introduction 85 7.5.2 Compactness (area density) 85

7.5.3 Heat transfer (heat flow density) 86

7.5.4 Discussion and conclusion 88

7.6 Conclusions 88

CHAPTER 8. GENERAL CONCLUSIONS

8.1 Retrospective view 8.2 Conclusions 8.3 Considerations 8.4 Recommendations 91 91 92 93

APPENDIX A - HEAT AND MASS TRANSFER CORRELATIONS OF THE SIMULATION PROGRAM

APPENDIX B - CALCULATION SCHEMES OF THE AHP COMPONENTS REFERENCES NOMEMCLATURE 97 99 103 109 CURRICULUM VITAE

₁₁₃

SUMMARY

This dissertation covers a theoretical and experimental study into the possible application of compact heat and mass exchangers (CHME) in a gas fired absorption heat pump (AHP) for domestic heating.

The framework of the study is defined by discussing the general principles and the research fields of the AHP. In addition the goals, the means and the set up of this research work are explained. (Chapter 1.)

The above-mentioned heat and mass exchangers are of the plate type. The space between the parallel and plain plates is filled up with corrugated plates of a certain height, the corrugation or finned plate.

The plain and finned plates are stacked and welded together. This gives a heat and mass exchanger which is very compact expressed by a high area density ( m2/ m3) . This leads to heat and mass transfer processes with small

temperature and concentration differences. (Chapter 2.)

For testing purposes a pilot plant was built using the above type of components in order to test their heat and/or mass transfer performance. Only the generator is of the shell and tube (SAT) type.

As the working pair CH30H - LiBr / ZnBr2 was chosen, with the alcohol as the

solvent and the salt mixture as the absorbent. This leads to sub-atmospheric working pressures with only solvent in the vapour phase. (Chapter 3.)

A literature survey has been conducted on working pairs for sorption systems in order to update the knowledge in this field and to select a new working pair for the pilot plant. This offers the possibility to verify and validate the simulation and the experimental results.

As possible new working pair R123a - DTG was selected and tested in the simulation program. For practical reasons experiments in the pilot plant are outside the scope of this research work. (Chapter 4.)

At the same time a computer program has been developed to simulate the test plant, based on heat and mass transfer correlations found in the literature, later to be replaced by correlations based on experimental results.

The program consists of three parts. The main program covers the overall calculation and iteration procedures. The component program contains the subroutines of the separate components, while the property program contains the thermodynamic and physical property data of several working pairs and cooling/heating media. (Chapter 5.)

Three series of experiments have been carried out, during which the input parameters were varied over a certain range.

The heating temperature of the evaporator was between 5 and 10°C, of the generator mostly 125°C due to the instability of the methanol in the mixture above that temperature).

The cooling temperature of the absorber and condenser was varied between 30 - 60"C over a mixture mass flow range of 25 - 125 g/s.

Different approaches have been adopted to interpret and explain the experimental results, the emphasis being on the absorber as the most

important and interesting component (simultaneous heat and mass transfer in a film flow).

The results have been used to find matching (heat) transfer correlations and to verify the film (heat transfer) and penetration (mass transfer) theory as adopted in the simulation program to match them by means of correction

factors. Also the SAT generator/absorber and the CHME absorber have been compared for their compactness (area density) and their heat transfer (heat flow density). (Chapter 7.)

Conclusions have been drawn concerning the possible application of the finned plate compact heat and/or mass exchangers in thermal sorption systems, while recommendations have been given for further research work. (Chapter 8.)

SAMENVATTING

Dit proefschrift doet verslag van het theoretische en experimentele onderzoek naar de mogelijke toepassing van zgn. " compact heat and mass exchangers (CHME)" in een gas-gestookte absorptie-warmtepomp (AWP) voor individuele woningverwarming.

Het kader wordt aangegeven middels een beschrijving van de werkings­

principes en van het huidige onderzoeksveld betreffende de AWP. Ook worden de doelstelling, de middelen en de aanpak van het onderzoek uiteengezet.

(Hoofdstuk 1.)

De bovengenoemde warmte- en stofwisselaars zijn van het plaat-type. De ruimte tussen de parallelle en vlakke platen is "gevuld" met een vervormde plaat, de zgn. "corrugation" of gevinde plaat.

De vlakke en gevinde platen worden gestapeld en tot een pakket gesoldeerd. Dit geeft een warmte- en stofwisselaar die zeer compact is en heeft een hoge oppervlaktedichtheid ( m2/ m3) . Dit geeft overdrachtsprocessen van warmte en

stof met kleine temperatuur- en concentratieverschillen. (Hoofdstuk 2.) Voor beproeving hiervan is een testopstelling gebouwd met dit type componenten om de warmte- en stofoverdracht te onderzoeken. Alleen de generator is van het "shell and tube (SAT)" type.

Als stofpaar is CH30H - Li Br / ZnBr2 gekozen met de alcohol als het

oplosmiddel (solvent) en het zoutmengsel als de absorbent. Dit geeft sub-atmosferische werkdrukken met alleen solvent in de dampfase. (Hoofdstuk 3.) Er is een literatuurstudie verricht naar stofparen voor sorptiesystemen. Dit om de kennis op dit gebied bij te houden en om een nieuw stofpaar voor de testopstelling te selecteren. Dit laatste biedt de mogelijkheid de simulatie en de experimentele resultaten te verifiëren. Als mogelijk nieuw stofpaar is R123a - DTG geselecteerd en beproefd in het simulatieprogramma. Experimenten in de testopstelling vallen helaas om practische redenen buiten dit

onderzoek. (Hoofdstuk 4.)

Tegelijkertijd is een computerprogramma ontwikkeld dat de testopstelling simuleert, werkend met warmte- en stofoverdrachtsrelaties bekend uit de literatuur. Later zullen deze vervangen worden door relaties komende uit de experimenten.

Het programma bestaat uit drie delen. Het hoofd-programma zorgt voor de "overall" berekeningen en iteratieprocedures. Het componenten-programma bevat de subroutines van de onderscheiden componenten, terwijl het stof-programma de thermodynamische en fysische stofgegevens van meerdere stofparen en verwarmings/koelmedia bevat. (Hoofdstuk 5.)

Er hebben drie series van experimenten plaatsgevonden waarbij een aantal invoerparameter over hun werkgebied gevarieerd zijn.

De verwarmingstemperatuur van de verdamper lag tussen de 5 en 10°C, die van de generator was meestal 125°C i.v.m. de instabiliteit van de alcohol in het mengsel boven deze temperatuur. De koelwatertemperatuur (abs./cond.) werd gevarieerd tussen de 30 en 60°C, de mengselmassastroom tussen de 25 en 125 g/s. (Hoofdstuk 6.)

Verschillende benaderingen hebben plaatsgevonden om de experimentele

resultaten te interpreteren. De nadruk lag daarbij op de absorber als de

meest belangrijke en interessante component (simultane warmte- en

stofover-dracht in een filmstroming).

De resultaten zijn gebruikt om correlaties voor de warmteoverdracht te

vinden, om de juistheid van het toepassen van de filmtheorie voor de

warmteoverdracht en van de penetratietheorie voor de stofoverdracht te

onderzoeken en om correctiefactoren voor deze theorieën te vinden.

Ook is een vergelijking gemaakt tussen de SAT generator/absorber en de CHME

absorber. Vergeleken zijn zowel de compactheid (oppervlaktedichtheid m

2

/m

3

)

als de warmteoverdracht (warmtestroomdichtheid W / m

2

) . (Hoofdstuk 7.)

Conclusies zijn getrokken voor de mogelijke toepassing van de gevinde plaat

CHME's voor thermische sorptiesystemen. Ook zijn een aantal aanbevelingen en

richtingen voor verder onderzoek aangegeven. (Hoofdstuk 8.)

CHAPTER 1. THE ABSORPTION HEAT PUMP

1.1 Introduction

This chapter is meant as an introduction to the absorption heat pump (AHP) in general and focuses on its most important aspects. This subject will be dealt with more specifically in the sections following this introduction.

In Section 1.2 an explanation is given of the working principle of the AHP, introducing the different heat and mass flows, the definition of the concentration of a mixture and the need for rectification.

In Section 1.3 the log P - l/T diagram, with the concentration as a parameter, of an absorption working pair (mixture / pure substances) is explained from a theoretical basis. The different processes in the AHP are presented in the corresponding diagram.

Section 1.4 defines some ratios concerning the performance of the AHP. These are the heat ratio, the coëfficiënt of performance and the circulation ratio.

Section 1.5 deals with the research field on AHP's. Apart from a distinction into theoretical and experimental research work, a distinction has been made into three different fields of research. Those fields are research regarding new working pairs, new system configurations and heat and mass transfer improvement in components.

The last section, Section 1.6, describes in the framework of the other sections mentioned above the research work reported in this dissertation. It focuses on the goals, the means and the set up of this AHP project.

1.2 Working principle

Almost all handbooks on absorption and on heat pumps contain a detailed description of the basic principle of the AHP, its components, the chosen workinq pairs and the processes of flow, heat and mass transfer (Kirn [KI], Bergmahs [BI], Stolk [SI]).

The figure on the next page shows a simple scheme of an AHP, with its components, the corresponding heat flows and temperature levels and the different mass flows (Figure 1.1).

In an AHP one is deal ing with a working pair, that is a solvent and an absorbent. The first is the volatile component, the second the less (or non-) volatile component. The concentration of a mixture, vapour or liquid, is defined as the weight fraction w, that is

w = kg solvent / kg mixture (1) In the AHP there are two mass flows (liquid/vapour), and both can be a

mixture of both components. That is first the fluid that is circulating in the cycle of the absorber, the heat exchanger and the generator, and

secondly the fluid/vapour that is flowing through the condenser and the evaporator, and then absorbed in the absorber and generated in the generator

(absorbed in and generated from the first fluid).

Depending on the boiling point difference AT between the solvent and the absorbent, the vapour generated in the generator can contain a certain amount of the absorbent, the less volatile component.

Qc, T c Mv. Wv

COND

RECT

Qe. Te

GENE

Qr. Tr

Qg. Tg Mp. Wp _

H.E.

Ph t

PI

EVAP

Mr, Wr Pp

H _ J _

ABSO

Qa. Ta Figure 1.1 Scheme of an absorption heat pump

If so, a rectification column between the generator and the condenser is necessary to ensure a weight fraction w of almost 1.0 in the vapour going to the condenser. The remaining absorbent in the vapour, with or without

rectification, will remain in the evaporator and must be transported to the absorber to maintain a steady state in the process.

In practice, a boiling point difference AT larger than 200 K will ensure a vapour weight fraction w close to 1.0.

So the first fluid is a relatively "rich" or "poor" (solvent) mixture of both components, the second fluid/vapour has a weight fraction w of 1.0 or very close to that, because, if rectification is necessary, it will never be a complete rectification of the vapour.

Let us now follow the process cycle, starting in the absorber (see Figure 1.1).

A rich mixture, coming from the absorber, pressure, through a heat exchanger, where the generator. By adding a heat flow Q at

vapour is generated from the rich mixture, which, now a poor mixture, returns via the heat exchanger, where it is lowered in temperature, and an expansion valve, where it is lowered in pressure, to the absorber. Now, the vapour enters, if necessary, the rectification column where it is cooled by removing a heat flow Q at a temperature T . The now rectified vapour passes to the condenser and the remaining "very poor" mixture returns to the

generator. The vapour, rejecting a heat flow Q at a "medium" temperature level T by condensation in the condenser, returns by another expansion valve, Feducing the pressure, to the evaporator where a heat flow Q is adapted at a "low" temperature level T by evaporation. Then the vapour (pure solvent) flows to the absorber and is absorbed by the poor mixture, rejecting a heat flow Q at a "medium" temperature level T . In the

following figure the different temperature and pressure lëvels (Figure 1.2) can be found with:

is pumped, causing a rise of it is raised in temperature, to

T = temperature [K] P = pressure [Pa] w = weight fraction [kg/kg] Q = heat flow [W]

Figure 1.2

Scheme of the AHP in a P - T diagram Ph !

P

PI Qe

* Qm

Qcf fQT

Qal Qg Te T Tm Tg

Under most conditions the heat flows in the absorber, in the condenser and in the rectification column are rejected at the same temperature level T_

so

'm'

VVV

Tm

The total heat flow output Qm is

m

Q = Q

Hm wa +

Q

C +

(2)

(3)

In the case of our working pair methanol - lithium bromide / zinc bromide (CH30H - LiBr/ZnBr2 (2:1 mol)), the methanol is the solvent. For this

working pair AT » 200 K, salt has a negligible vapour pressure holds for the well-known working pair lithium bromide/water (H20

So in both cases rectification is not necessary.

The mixture of our working pair has a weight fraction w between and 0.40. In the case of the also well-known working pair NH3 - H20 ,

solvent is ammonia, AT = 135 K, so rectification is a necessity.

The same - LiBr).

0.28 the

1.3 The log P - l/T diagram

When deal ing with a working pair, one has one extra variable, namely the weight fraction w. In other words, the liquid mixture possesses two degrees of freedom at liquid-vapour equilibrium, so w = w(P,T), P = P(w,T) and T = T(w,P).

Derived from the well-known Clausius - Clapeyron law, the following relation between the absolute vapour pressure P and the absolute temperature T, known as the Antoine equation, can be written

with r = enthalpy of evaporation/condensation [J/kg] 1 = enthalpy of absorption/desorption [J/kg] a,a'= constant

b,b'= constant

With this in hand one can draw for this type of mixture a log P - l/T diagram with the weight fraction w as parameter. One should keep in mind

that every point in this diagram describes an equilibrium situation, so a process can not be drawn in the diagram, only the two points between which it takes place and the direction. The "processes" in the AHP are drawn in the following log P - l/T diagram (Figure 1.3).

Figure 1.3

The log P - l/T diagram

for an AHP with the weight

fraction w as parameter

In P

1/T

In this diagram two important simplifications have been made, namely that the enthalpy of evaporation/condensation and the enthalpy of absorption/ generation are independent of temperature and weight fraction. Nevertheless this diagram gives a good overview and presentation of the processes taking place in the AHP.

1.4. The heat ratio, the coëfficiënt of performance and the circulation ratio

If one considers the AHP as a reversible one, one can, with the ingoing outgoing heat flows at three different temperature levels (Figure 1.2), define a so called heat ratio £, based on the Carnot efficiency:

and

« ■ <Y

T

e» I \ ■ \ I <V

T

e>

(6)

The heat flow Q is adapted in the evaporator at the temperature level T , the ambient temperature. Therefore this heat flow is not to be accountedefor

in Equation (6).

Furthermore, the process in the generator and in the absorber take place at a temperature range, here a mean temperature is assumed for both.

With this in hand, one can also write for the heat ratio, based on the "ins" and "outs" and the enthalpies:

e - [(r + l) + r] / (r + 1) (6a) and with

# - 1 / r (7) C - ( 2 + # ) / ( l + # ) (6b)

Most of the working pairs have a positive 0, so for a (single stage) AHP the heat ratio will never exceed 2.

As one can see the heat ratio is a pure theoretical parameter, also some assumptions had to be made, as well as some simplifications.

A more practical parameter, mostly based on the output of experiments, is the so-called coëfficiënt of performance, the COP, which is defined as follows:

COP = energy output / energy input (8) For the AHP that leads to

with:

COP = ( Qa + Qc + Qr) / ( Qg + Pp ) (8a)

Q = heat flow from absorber [W] Q = heat flow from condenser [W] Q = heat flow to generator [W] Q^ = heat flow from rectif. column [W]

P = pumping energy to pump [W]

Since P « Q , the pumping energy P is often left out from Equation (8a). This coëfficiënt will be used later on in the experiments with the computer simulation model and program and with the test plant as a main parameter for the performance of the AHP.

Another practical parameter is the circulation ratio f, the mass flow ration of the rich mixture and the vapour leaving the rectification column:

f = Mr / Ms (9)

with: f = circulation ratio [kg/kg]

M = mass flow [kg/s] r = rich, p = poor, s = solvent

By using two mass balances, of the total mass and of the mass of the solvent, the circulation ratio can be expressed in terms of the weight fraction w:

total mass balance M + M = M (10) mass balance solvent w • M + M = w • M (11)

leading to: f • (1 - W ) / (wr - w ) (9a)

Together, the total heat production Q , the coëfficiënt of performance COP and the circulation ratio f form the fflain output parameters of the AHP.

1.5 Research work on absorption heat pumps

The ongoing research activities and those which took place in the past in the field of the AHP can, generally speaking, be divided into three research categories:

a. research on (new) working pairs,

b. research on new sorption system configurations and c. research on heat and mass transfer (components).

Within these fields one can also make a distinction in theoretical and experimental research activities.

It is not a surprise that there always is one main goal behind the direct goals of the research work: saving energy.

Looking at the last five years, it seems that more and more the

research work on sorption system is focussed on new system configurations. Mostly by means of a theoretical approach, that is by using computer

simulation techniques. It is in this research field that researchers expect the most promising results concerning energy saving.

In the research work on (new) working pairs for sorption systems, only a very few promising ones have been found up until now and they are still in the experimental stage. The well-known working pairs NH3-H20 and H20-LiBr

are still favourable because of the experience and reliability of these working pairs, also taking their disadvantages into account.

Research work on heat transfer, with or without change of phase, takes place extensively. This is not so when also mass transfer takes place.

So in components with simultaneous heat and mass transfer, which is the basic principle of sorption, not so much results can be found.

Although this research work is mainly focussed on components, on

exchangers, for simultaneous heat and mass transfer, for heat transfer with or without a change of phase, all three research fields will be introduced briefly.

1.5.1 Working pairs for an absorption heat pump

The well-known working pairs NH3 - H20 and H20 - LiBr are the most applied

working pairs in an AHP, but a world-wide research took place at technical universities and research institutes to find better working pairs, or working pairs that could meet the disadvantages of the above mentioned working pairs.

In an AHP for example, the working pair NH3 - H20 has high working

pressures and the necessity of rectification. Limitations in the use of the working pair H20 - LiBr are the solubility (danger of crystallization) and

the evaporation temperature (danger of freezing).

The following list gives an overview of the most important selection criteria for a working pair for an AHP. Of cause some criteria are more important than others or are in a way depending on one or more other criteria.

1. the heat of evaporation r [J/kg] of the sol vent,

2. the maximum working pressure P, [bar.kPa] of the solvent, 3. the circulation ratio f [kg/kg]'oT the mass flows,

4. the energy consumption P [W] of the mixture pump,

6. the critical temperature T [K] of the solvent and the absorbent, 7. the boiling point differente AT [K] between the solvent and the

absorbent,

8. the solubility in the desired operation field (no crystallization), 9. the chemical stability at the maximum desired temperature,

10. the corrosion to construction materials, 11. the toxity and

12. the availability and the costs.

At the Technical University of Essen, West Germany, a great effort has been made in the measurement of the thermodynamic properties of and the

composition of the P - T and h - w diagram for promising working pairs. Based on the list above, they selected the following criteria: 1. the heat of evaporation at 0"C, r0 [J/kg],

2. the pressure of the solvent at 50°C, P ™ [bar], 3. the circulation ratio f [kg/kg],

4. the pumping energy factor N :

Mp- (f • AP) / (rQ . pr) (12),

with: AP = pressure difference over the pump [kPa]

p - density [kg/m3]

5. the heat exchange factor N. :

Nh = ((f-D • cp • AT) / rQ (13),

with: AT = temperature difference absorber and generator [K]

6. the boiling point difference AT [K], the necessity of rectification and 7. the toxity

In the fundamental experimental research to new working pairs, that is the determination of the thermodynamic and physical properties, a great effort is put over the last twenty years to meet the disadvantages of and replace the well-known working pairs like NH3 - H20 and H20 - LiBr. Although

successes are made in this direction, the disadvantages of the new working pairs and the advantages of the conventional working pairs, leads to a tendency to choose for reliability and experience, so to conventional working pairs.

Only a very few new working pairs have been successfully, that is, are applied in experimental test plants of a sorption system. But the

conventional working pairs are still favoured in industrial applications, they are the only systems in practical use today.

The problem is that most of the new and promising working pairs fail on one or more criteria. In most cases corrosion and/ore chemical stability is the bottleneck, furthermore the toxity seems to play an important role.

Around 1980 methanol (CH

3

0H) seemed to be the solvent of the future, in

combination with LiBr or a mixture of LiBr and ZnBr

2

as the absorbent.

Unfortunately the methanol itself and the methanol/salt mixture showed to be

chemically unstable above 110 - 130°C, as detected by Koebel [K2].

That meant that only a maximum temperature (in the generator) of about 120"C

is allowed and, with an evaporation temperature of about 0°C, the useful

heat can only be produced at about 40 - 50°C. This drastically diminishes

the chances for a direct gas-fired domestic AHP heating system.

Until now, one of the most promising solvents (working fluids) seems to

be the 2,2,2-trifluorethanol (TFE, trifluorethanol, CF

3

CH

2

0H). The

thermodynamic properties are extensively investigated by Bokelmann

(TU-Essen)[B2, B3] and Girsberger (TU-Bern) [Gl, G2].

Working pairs with TFE as the solvent have already been tested in some AHP

or AHT pilot plants. Berghmans [B4] tested the working pair TFE - Chinoline

(C9H7N) in an absorption heat transformer (AHT) with a heat output of 275

kW. Nakayama [NI] has developed a sorption system (AHT) with the working

pair TFE - N-methyl 2-Pyrrolidon (NMP, C

5

H

9

N0).

Bokelmann [B3] investigated the possibilities of several working pair with

TFE of which the TFE - NMP showed to be the most promising one. Since with

NMP there is the need of rectification, he later changed to 2-Pyrrolidon

(Pyr, C

4

H

7

N0) [B5].

This last working pair, TFE - 2-Pyrrolidon, will now be tested in an

AHT test plant (heat output 10 kW) at the Laboratory of Refrigeration

Engineering at the Technical University of Delft, the Netherlands [Wl].

1.5.2 System confiqurations

As the absorption heat pump AHP is in fact developed from the absorption

cooling machine ACM, the absorption heat transformer AHT is developed from

the AHP. The following figure (Figure 1.4) shows the different temperature

levels of these three basic sorption systems.

New configuration of sorption systems are more or less based on these

three systems.

Figure 1.4

Temperature levels >

in the ACM, the

AHP and the AHT ACM AHP AHT

The first to mention is the two or multi-stage configurations to achieve

higher temperature differences and/or a better performance (COP) (Hobling

[Hl], Alefeld [Al], Ziegler [Zl]. Although some experimental research has

taken place with test plants, most is done by means of computer simulation

For the ACM and the AHP a higher temperature lift (double-lift) or a

higher COP (doublé effect) can be achieved. For the AHT a higher COP at a

smaller temperature lift (doublé effect) or a higher temperature lift at a

lower COP (doublé lift) can be achieved [Zl]. Figure 1.5 shows a two-stage

AHP system.

double-lift

(temperature)

_{Ore. Tre}

dg. Tg Ode. Tde Ppa -« Qa. Ta *•

b. doublé effect

(COP)

Qg. Tg Qc, Tc

Figure 1.5

A two-stage absorption

heat pump system

Qe, Te

In practice it seemed that more than two-stage is not yet feasible. For

industrial application more and more two-stage configurations are found

because of the increasing possibilities to reduce the investment costs.

An also possible configuration is the resorption heat pump or heat

transformer (RHP / RHT). In the AHP / AHT the evaporator and the condenser

circuit is replaced by a solution circuit with a desorber (a generator at

low pressure and temperature) and a resorber (an absorber at high pressure

and temperature). In fact the complete name should be the absorption /

resorption heat pump. So the "evaporator" and the "condenser" are now

working with gliding temperature differences. Figure 1.6 is showing a

resorption heat pump system.

The choice between an absorption or a resorption system is not easy

(Westra [W2], v/d Ree [Rl], Baehr [B6]), because the differences in

performance are small. The tendency is that the profits from the gliding

temperatures are smaller than the losses because of the extra investment

costs.

Also in view of the energy efficiency the absorption system seems to be

favourable. For the adaptability to changing process temperatures, a

resorption system seems to be better.

Ore. Tre

Figure 1.6

A (absorption/) resorption

heat pump system

Qde. Tde RESO

-1

H.E.r

n

RFNF

L

KEa

-J

P p r Ph

t PI

DESO (J

T

ABSO O» Tg Ppa Qa. Ta

An other new configuration is a combination of the RHP / RHT and the

compression cooling machine CCM and the compression heat pump CHP. That is

the compression heat pump with solution circuit. Instead of the evaporator

and the condenser a solution circuit with an absorber and a generator is

used. Figure 1.7 shows such a system.

Ore. Tre

RESO

Figure 1.7

KE.

A compression heat

pump system with a

solution circuit

Qde. Tde

Pp

Ph

PI

Pcomp

DESO

As with the resorption system, there is the possibility to work with gliding

temperature differences (Ahlby [A2]) to achieve a higher COP. Another name

for such a configuration is a compression/absorption system.

1.5.3 Heat and mass transfer

In sorption (absorption, resorption, desorption, etc.) the basic principle

is that in the components for these processes, simultaneous heat and mass

transfer is taking place, independent of the type of component (bubble,

film, spray, e t c ) .

But on micro-scale, where one is interested in the temperature and weight

fraction distribution (boundary layers), it is very difficult, if not

impossible, to measure temperature and weight fraction.

To meet this, simulation models are developed based on the different

theories on heat and mass transfer (film theory, penetration theory, etc.)

like v/d Wekken c.s. [W3] and Grossman [G3] did. But for a better

understanding and improving of the sorption processes, this simplification

is necessary.

In this research work there were used to predict the right surface

configuration for the different processes. Key word in this procedure is the

area density 0, the transfer surface area [m

2

] per volume [m

3

] of the

component.

1.6 This research work

1.6.1 Goals

The scope of the research reported here is a theoretical and experimental

study into the possibilities of the application of so called compact heat

and mass exchangers (CHME) in gas fired AHP's for domestic heating.

This heat and mass exchanger, the CHME, is of the plate type. The space

between the parallel and plain plates are filled with corrugated plates of

some height. This is what is called the corrugation. The plain and

corrugated plates are stacked and, with headers and pipes, brazed or welded

together. This qives a heat and mass exchanger which is very compact and has

more than 700 m

2

transfer surface per m

3

. This leads to heat and mass

transfer with small temperature and concentration differences.

Besides this more qualitative aspects, more quantitative information is

needed to say more about the behaviour and performance of this type of heat

and mass exchanger. This leads to a theoretical and experimental research

into the application of this type as a component of an AHP and well as:

- absorber and generator: simultaneous heat and mass transfer

- evaporator and condenser: heat transfer with change of phase

- mixture-mixture heat exchanger: heat transfer

Next to that, the AHP as a whole - its behaviour and its performance at all

desired working conditions - will be subject of this research.

In that frame work an AHP with heat and/or mass transfer components

built as compact heat and/or mass exchangers (except for the generator,

which is of the shell and tube type) was tested with the working pair

lithium bromide LiBr / zinc bromide ZnBr

2

(2:1) and methanol CH

3

OH, so a

salt / alcohol mixture with alcohol as the solvent. The thermodynamic

properties of this working pair are well known and it is extensively tested

in an earlier AHP test plant by Iedema [II] and Saurwalt [S2].

Therefore this working pair is suited to test the compact heat and/or

mass exchanger components in a new AHP pilot plant, although one of the most

important and limiting disadvantages of this working pair is the fact that

the methanol in the mixture is not stable above temperatures of 120°C (393K)

and is breaking up, a process which is not reversible.

1.6.2 Tools

The starting point of this research study was the research work of Iedema, especially his dissertation "The Absorption Heat Pump" [II].

In this one can find the first set up for the development of an AHP for domestic heating.

Based on the thermodynamic properties, derived experimentally and theoretically by the fundamental equations of state, of the until that time known working pairs, the most suitable one was chosen, LiBr/ZnBr2 (2:1)

-CH3OH.

With that in hand a computer simulation model and program was developed by Bakker [B7] to simulate an AHP for domestic heating with this working pair. In this program all the components were put in separated subroutines and were of the shell and tube type, with straight or wound tubes. Only the mixture heat exchanger was of the plate type.

The complete program was run through until all the components itself and the AHP as a whole were in balance. The main input variables were the ambient temperature and the chosen control strategy (mixture mass flow and additional heating).

Furthermore by means of a literature study on flow hydrodynamics, heat and/or mass transfer, all concerning the falling film flow around tubes, a theoretical model was developed for the absorber, consisting of rows of parallel tubes, and for the occurring processes of flow, heat and mass transfer. It should be emphasized that this dissertation was mainly

concentrating on the absorber as the most important and limiting component

of the AHP.

Finally an AHP pilot plant was built to test the absorber and the heat pump as a whole. Also here the components were of the shell and tube type, only the mixture heat exchanger was of the plate type.

For a better heat and mass transfer and a more compact construction (domestic heating !) the step is made to the above mentioned compact heat and mass exchangers, CHME.

So for this research, there are three starting points:

a. the dissertation of Iedema [II], mainly the chapters 3 until 7 deal ing with heat and/or mass transfer, the absorber model and the

simulation of and the experiments with an AHP,

b. the computer simulation model and program of an AHP for domestic heating with a salt/alcohol mixture [B7],

c. an AHP pilot plant with components built with so called compact heat and/or mass exchangers, only the generator is still of the shell and tube type.

1.6.3 Set up

This work, with the above mentioned three starting points, has also three basic elements, namely simulation/modelling, experiments and literature.

The next chapters will contain the results of this research work, realized along the above mentioned lines.

CHAPTER 2. COMPACT HEAT AND MASS EXCHANGERS

2.1 Definition

Compact heat and/or mass exchangers have a compact transfer surface, with a

area density greater than 700 nr/m

3

, a somewhat arbitrary value. Heat and/or

mass exchanger stands for the transfer surface of heat and mass, while

compact stands for not only compact itself, but also for enhanced.

In short, a compact heat and/or mass exchanger is an exchanger for heat

and/or mass of a compact construction with many "extra" transfer surface and

a relatively small volume.

2.2 Compact and enhanced transfer surface

Concerning the heat transfer, one can define the amount of heat which is

transferred from one side of the exchanger to the other side as follows:

Q = K • A • AT (14)

with: Q = heat transfer rate [W]

K = overall heat transfer coëfficiënt [W/m

2

.K]

A = total heat transfer surface area [m

2

]

AT= mean temperature difference [K]

Furthermore one can define for the heat exchanger

7 = Q / AT (15)

with: 7 = specific heat transfer rate [W/K]

and

0 = A / V (16)

with:

f} = area density [m

2

/m

3

]

V = total volume [m

3

]

For exchangers of the plate type, this leads to

h =

A

c /

V

c

or

*h =

A

h /

V

h (

16a

>

with: c = cold side, h = hot side

and for exchangers of the shell and tube type, this leads to

0

t

= A

t

/ V

t

(16b)

with: t = total

An clear overview of the different types of exchangers concerning their area

density is derived from Shah [9] and shown in Figure 2.1.

-COKPACTNESS?" * KATTER OF DEÜRf-E

o o o o o o

o o o o o o

o o o o o o o o o o 1 o o o o o o o o o o o o o o o o o o o o o I ^ Q O _ o o o o o o o o y \-"'''"'" • • • • " > " " • •>>"

KH

„-O O

O-r

O I O O ^

Gai Turbin. Rotary iHegenccitoi Buun : ...

D

Automotivr f U d i a t o r s

Matrix Types, H.r* Screen Sphere Bed, Corrugated Sheets

For X X , ■ 1.B8,

SL

S t r i p - F i n and Louv^red-Fin H.E. ■• E ■ j p Jnd O ■ O.B!]

P U i n T u b u U r , Shell-and-Tub« H.E. COMPACT SURFACES

20 10

llydrdullc DiaawtU DL

5 2

T — l l l i l i l

200 500 1 (

Heat translVi BürfacÉ- * I , I I I I I I 60 100

1 r

2000 I M U I . I I I . 5000 10 2 1x10

Figure 2.1 The area density /? for different types of exchangers

from Shah [9]

Figure 2.2 A stack of parallel plain and corrugated plates

plain triangular fin plain rectangular fin wavy fin

offset strip fin round perforated fin pin fins

The goal is to achieve a great specific heat transfer rate in combination with a small mass and volume of the exchanger by means of the use of compact

surfaces. This most of the time leads also to a greater overall heat transfer coëfficiënt K which itself also leads to a smaller volume. This compact construction is strong and stable with small wal! thicknesses. This also reduces the volume and to a smaller content the mass of the exchanger.

With enhanced surfaces is meant to achieve a greater overall heat transfer coëfficiënt compared to the not enhanced plain surfaces. So it deals with the factor K • A (= y ) , achieving an enlargement of K and/or A.

This can be done by

a. adding extended surface to the prime surface (K and A greater) b. adding turbulators to the surface (K greater)

c. reducing the flow passages (hydraulic diam.) (K and A greater) As one can see, a. and c. also increase the transfer surface area. Compact and enhanced transfer surfaces characterizes more the category of compact exchangers than the earlier mentioned area density of 700 m2/m3. Clearly is

shown that in most cases compact and enhanced appear together.

This all offers the possibility to apply this type of exchanger with an extensive choice in type, geometry and area density of the surface. So a great flexibility in the choice of surface on both sides and a reduction in the total mass and volume can be obtained.

This also offers the possibility of automated production techniques, certainly in the case of the plate fin type exchangers as used in this research work and shown in Figure 2.2. This all can lead to a competitive price of this type of exchanger.

2.3 Application

2.3.1 Field of application

The most important application of compact heat and/or mass transfer surfaces concerning the type of mass flow, is on the gas side in heat exchanging processes (in gas/gas, gas/liquid, gas/condensing,evaporating liquids). To a much smaller extend they are used for applications for two phase flow or on the liquid side.

It is obvious that for a certain rate of transferred heat a much larger transfer surface is needed on the gas side than on the liquid side because of a normally 10 - 100 times smaller heat transfer coëfficiënt on that gas side. Unfortunately the enhancement of transfer surface by means of

corrugations or turbulators gives a much greater pressure drop on the gas side. Therefore the adding of that type of turbulators is limited.

Application in the aircraft and automobile industry was the starting point of the development, foliowed by application as heat transfer

components in installations [S4]. The application in sorption systems is rather new, up until now hardly no literature is found on that, except for Minkus [Ml].

2.3.2 Corrugations

If one limits oneself to compact heat and/or mass exchangers consisting of a stack of parallel plain and corrugated plates, one can make the following classification for the type of corrugation [S3] (Figure 2.3):

- plain fins (rectangular, triangular, square)

- wavy fins

- offset strip fins

- louvered fins

- perforated fins

- pin fins

2.3.3 Construction materials

Construction materials are mainly aluminium and copper while stainless steel

and other corrosion resistant materials are only scarcely used. This is

mainly caused by the difficult manufacturing and the brazing of the

corrugated plates of this materials (AKG [A3], Trane Company [Tl]).

2.4 Characterization

2.4.1 Identification

To identify the performance and quality of each type of surface (here

corrugated plate) two characteristic properties are used, one for the heat

transfer, j , also called the Colburn factor, and one for the friction, f,

also called the Fanning friction factor [K3]. Both are non-dimensional and

defined as follows:

j = St • P r

2 / 3

(17)

f = 2 • to /

(p • v

2

) (18)

with: St = Stanton number (St = Nu/(Re«Pr))

Pr = Prandtl number

Nu = Nusselt number

Re = Reynolds number

to = surface shear stress [kg/m.s

2

]

p = mass density [kg/m

3

]

v = velocity [m/s]

When the thermodynamic properties of the flow medium are known, one can

define j and f for a given geometry of the corrugation as a function of the

Reynolds number Re.

An other form of presentation which is also commonly used in literature

is the following one:

and

Nu = a • Re

b

• Pr

c

(here c=l/3) (19)

AP = f • d • 1/2 •

p • v

2

(20)

with: a, b, c, d = form factors (constant) [-]

AP = pressure drop [N/m

2

]

So in this way the factors j and f, or Nu and AP, are commonly used to

express the performance of the type of transfer surface (corrugated plate).

2.4.2 Analvtical solutions

In Shah and London [S5] one can find a great amount of analytical solutions,

that is for simple geometrical configurations like triangular, rectangular

and circular channels. It is restricted to fully thermal and

hydro-dynamically developed laminar flow, that is fully developed temperature and

velocity profiles. This gives a constant Nusselt number, independent of the

Reynolds and Prandtl number, only depending on the geometry.

For three sets of thermal boundary conditions the Nusselt numbers are given:

1. constant wall temperature,

2. constant wall temperature and heat flow in the axial direction and

3. constant heat flow in the axial and peripheral direction.

The characteristic length in the Nusselt number and in the earlier mentioned

Reynolds number is the so called hydraulic diameter d

d

h a

4 • A

r

• 1

h'

(21)

with:

d, = hydraulic diameter [m]

A

f

f flow cross sectional area [m

2

]

A = total heat transfer area [m

2

]

1 = flow length [m]

For most of the geometries one can write:

h = - Q - f

with:

0 = wetted perimeter [m]

(21a)

For a better understanding of the above mentioned formulas, this means for a

rectangular channel (see Figure 2.4) :

d

h - i

A

r

• 1 _ 4 • (w • hl « 1 2 • w • h

A 2 • (w + h) • 1 w + h

(21b)

with:

w = flow passage width [m]

h = flow passage height [m]

1 = flow passage lenght [m]

If w » h than d

h

« 2

Figure 2.4

Offset strip fin

corrugated plate

t = fin thickness [m]

x = uninterrupted fin length [m]

2.4.3 Experimental results

2.4.3.1 Introduction

The type of heat and/or mass exchangers that is used in the earlier

mentioned absorption heat pump (AHP) pilot plant is of the plate fin type.

The transfer surface is corrugated and of the offset strip fin type.

Figure 2.4 shows an example of this type of corrugation.

In the following the limitation is made to heat exchangers of the plate

type and to heat transfer. This because most of the available literature

deals with heat transfer, and the exchangers in the pilot plant are of the

(corrugated) plate type as shown Figure 2.4.

2.4.3.2 The work of Kavs and London

A Standard reference in this field is the book "Compact Heat Exchangers",

written by Kays and London [K3], It gives a thorough and detailed survey on

compact heat exchangers, not only concerning the geometry, but also a very

large amount of experimental results.

Not only for transfer surfaces of simple geometry, but also for

very

complicated geometries experimental results are given for tubes and plates.

For all these types data are given for the factors j and f as well as a

clear indication of the geometry. Almost all the experiments took place with

an air flow through the corrugated passages and a (condensing) steam flow

through the other passages. For the overall heat transfer coëfficiënt K in

Equation (14) one can write:

K = [ 1

/a

Q

+ d y ^ + 1

/a

s

V

1

(22)

with:

a- heat transfer coëfficiënt on the air side [W/m

2

.K]

d = plate thickness of the wall [m]

X » thermal conductivity of the wall [W/m.K]

a = heat transfer coëfficiënt on the steam side [W/m

2

.K]

For this kind of heat exchange it is allowed to say:

a »

a and X/d »

Ö L

so: K * a (22a)

s o ' o o

2.4.3.3 Rectanqular offset strip fin surfaces

Since the corrugated surfaces of the components of the pilot plant are of

the rectangular offset strip fin type, the focus will be on that type of

corrugated surfaces.

Wieting [W4] has gathered all the experimental results for rectangular

offset strip fin configurations from [K3] and [S6] and correlated the

experimental heat transfer and flow friction data over two Reynolds number

ranges, that is for Re < 1000 (laminar) and Re > 2000 (turbulent).

A clearly defined transitional Re, Re*, was not found. So to minimize

the effect of the transitional Re, occurring basically between 1000 < Re* <

2000, on the correlation, this range was excluded.

To find these correlation for heat transfer (and for flow friction an

analogous one, that is f) the following non-dimensional functional relation

was assumed:

j = a • (x/d

h

)

b

- (t/d

h

)

c

. (w/h)

d

- Re

e

(23)

(a, b, c, d, e en d are unknown coefficients.)

The Reynolds number is based on the hydraulic diameter d. . The ranges of the

indicated variables are as follows:

0.7 < x/d. < 5.6

0.03 < t/d" < 0.166

0.162 < w/h

n

< 1.196

0.65 < d. < 3.41

100 < Rë < 10000

Furthermore, the area density

p is in the range 1000 < fi < 3000.

Leaving out the flow friction factor, Wieting found the following

correlations, based on the available heat transfer data of 22 offset strip

fin surfaces:

j = 0.483 • (x/d,)-°-

162

. ( w / h ) "

0

'

1 8 4

. R e "

0

'

5 3 5

,

for Re < 1000

n

(24)

j . 0.242 • (x/d.)-°-

322

. ( t / d . )

0

"

0 8 9

. R e "

0

"

3 6 8

,

for Re > 2000

n n

(25)

This with j from Equation (17) and

Nu = (

a • d

h

) /

X (26)

Re - (

fi

. v • d

h

) /

r\ (27)

Pr = ( t) . c ) / X (28)

with:

r? = dynamic viscosity coëfficiënt [kg/m.s]

The overall discrepancy between the correlations and the experimental

results is within 10 %.

For the transitional Re Wieting derived from Equation (24) and (25) the

following equation for Re*:

Re* = 61.9 • ( x / d . )

0

"

9 5 2

. (t/d.)"

0

-

5 3

. ( w / h )

- 1

-

1

(29)

for 1000 < Re* <

n

2000

n

So if Re < Re*, than use Equation (24) and when Re > Re*, than use Equation

(25) in the transitional range.

For any surface geometry of the corrugated plate one can with this

calculate the Colburn modulus j as a function of the Reynold number Re.

In the same way Wieting derived equations for the friction factor f.

If one correlates Equation (24) and (25) for all the 22 offset strip fin

surfaces, one can find an "average" correlation for this type of surface for

all geometries:

n

„,.

j = 0.487 • R e "

u

-

D J D

for Re < Re* (30)

j = 0.149 • R e

- 0

'

3 6 8

for Re > Re* (31)

For the use of the Equation (24) and (25) and even more of the Equation (30) and (31) one should be careful in extrapolating data for strip fin

geometries that have geometrical parameters outside the range of those for the correlations. This one is strictly based on a limited amount of reported test data. Also these correlations may be applicable only for air or gas as the working fluid (Pr « 1 ) .

Mochizuki and Yagi [M2] have done experimental research with aluminium test cores of plate fin type heat exchangers with (offset) strip fins, also using air as the test fluid and condensing steam as the other medium.

Seven types of strip fin surfaces were tested, with a constant height h (10 mm) and fin thickness t (0.2 m m ) . They were mainly interested in the influence of the fin spacing (width) w and the fin length x.

The range of the parameters is here in the same range as those belonging to the data Wieting [W4] used in Equation (23), except for the hydraulic

diameter dh, here 3.04 and 4.35 mm. The corresponding area densities ,8 are

700 and 1000 m2/ni3.

They also presented their results in the form of the Colburn factor j for the heat transfer and the Fanning factor f for the flow friction as a function of the Reynolds number Re, but in a different way than Wieting did. Here the wall surfaces of the flow passages are interrupted in the flow direction and the process is repeating itself since the air flows along the very short strip surface and then separates at the trailing edge of the strip. Thus the boundary layer is never able to become thick. Therefore, if the pressure loss is considered to consist of the two effects of form drag of the fin and a friction drag of the fin surface owing to the viscosity of the fluid, then the skin friction factor f may be expressed in the following

form: Q

f = a + b • Re 1 , ü (32)

and a similar expression for the Colburn factor j :

j = c + d • R e "1 , 0 (33)

For each type of strip fin surface they defined a set of a, b, c and d. For comparison the following averaged correlation for the Colburn factor was derived from their data:

j = 4.72 • 10"3+ 10.06 • R e "1 - 0 for 1000 < Re < 8000 (34)

For comparison with the correlations of Wieting for the Reynolds number range 2000 - 8000, Equation (34) was transformed into the following one:

j * 0.135 • R e "0 - 3 5 0 for 2000 < Re < 8000 (35)

Within this range the maximum deviation in j from Equation (31) is less than 6 %.

2.5 Discussion

Shah and Webb [S3] have discussed both the results of Wieting and Mochizuki and Yagi.

They started with a theoretical solution. For the heat transfer they took the Pohlhausen laminar boundary layer solution for a flat plate of "plate length" x [S7]:

j = 0.664 • R e ^0 - 5 (36)

and f o r the f r i c t i o n the modified Blasius laminar boundary layer s o l u t i o n f o r a f l a t p l a t e using the form drag associated with the leading blunt edge of the s t r i p f i n [S7]:

f =(Cd • to) / ( 2 • x) + 1.328 • R ex°5 (37)

with: C . = form drag coëfficiënt [-]

In order to indicate and compare the performance of offset strip fin surfaces, one can use the factor j/f as an indication.

Based on the Reynolds analogy for flow over a flat plate, in the absence of form drag ( C. = 0 ) , the factor j/f should be 0.5 (for Pr*l). Since the contribution of the form drag is of the same order of magnitude as the skin friction drag for such an interrupted surface, the factor j/f will be about 0.25.

If one takes the correlations of Wieting of Equation (30) and (31), one will find that for the ranges Re < 1000 and Re > 2000 the factor j/f is

smaller than 0.275. Even in the transitional range the factor j/f does not exceed 0.30.

Mochizuki and Yagi used the factor j/f to determine the influence of the geometrical factor x/w and the Reynolds number Re. They found, with a negligibly small dependence on the Reynold number Re, for the factor j/f that 0.2 < j/f < 0.6.

Shah and Webb have their doubts about the results of Mochizuki and Yagi because the factor j/f > 0.3 and stated that "published data for strip fins are questionable if j/f > 0.3. All the measurements and possible sources of flow leaks and heat losses must be checked thoroughly for all those basic data having j/f > 0.3 for strip fins."

Dubrovskii and Fedotova [Dl], [D2] investigated an offset strip fin surface, not a rectangular one but with a slit form, with d, = 2.9 mm and £

= 1280 m2/m3. n

They found the following expressions for j and f:

j = 0.090 • R e "0 - 3 0 for 800 < Re < 3250 (38)

f = 1.590 • R e "0 , 2 7 for 1500 < Re < 3250 (39)

As one can see, in this range the factor j/f is almost independent of that Reynolds number and the factor j/f = 0.045. Compared with earlier results this is rather low.

Also attempts have been made to predict the factors j and f numerically by a finite difference method for the offset strip fins considering the laminar boilndary layer on each strip fin. In [S3] Sparrow [S8] and Patankar [PI] are mentioned. They compared the numerical results with a strip fin surface from Kays and London [K3] and found a reasonable agreement for the factor f, but the predicted j factors were about 100 % higher. The predicted slopes of both j and f versus Re curves were steeper than the test data.

One of the major unpredictable factors, mentioned in both [S3] and [K3], is the existence of small burrs at the leading and trailing edges of the fin during its formation by a shearing operation. Fins of this type are generally constructed by a machine-cutting process that inevitably leaves a slightly bent and grazed fin edge that differs depending upon the fin

material and character of the cutting tooi. These burrs increase the

effective plate thickness, causing increased form drag. This factor can not be taken into account accurately in the numerical solutions, nor can the influence on the experimental data be estimated accurately. Therefore, the possibie existence of burrs causes uncertainty in the correlations or in the comparison of predicted values with data. Briggs and London [B8] have paid some attention to this point.

So far the theoretical, numerical and experimental information found on the performance of rectangular offset strip fin surfaces in literature.

2.6 Application in sorption svstems

In the last section of this chapter on compact heat and mass exchangers, a global comparison will be made between the application as described in the sections before and the application in a sorption system, in particular in the AHP test plant. The main important differences are pointed out.

First of all should be mentioned that one has to do with simultaneous heat and/or mass transfer and/or with a change of phase, not only with heat transfer. In literature only information is found concerning pure heat transfer.

In the cases of the analytical solutions and the experimental results the flow was in the direction of the "most open" side of the corrugated passages. In the exchangers of the AHP the flows on both sides - both sides have corrugated passages - are in the direction of the "most closed" side of the corrugated passages. Because the strip fins are offset, the passaqes are not completely blocked.

The working fluid flowing through the corrugated passages was air, with on the other side condensing steam. In the AHP, in the secondary circuits

(for cooling or heating) water or liquid methanol were used as media while in the primary circuits lithium bromide/ zinc bromide - methanol (the

absorbent or mixture) and methanol (the sol vent), liquid and/or vapour, were the media.

To have a rough indication of the thermodynamic properties of these media, in comparison with those of air, one can find in Figure 2.5 the Prandtl number Pr, the kinematic viscosity v and the thermal conductivity \

over a certain temperature range for the different media. By this one can estimate the possibie influences on the heat transfer, keeping in mind the equations for the Nusselt number, the Reynolds number and the Prandtl number

(Equation (26), (27) and (28)) and the heat transfer correlation of Equation (19).

100 5 0 - 20- 10- 0.5-0.2 PrandN X-10*(W/m-K] V-10* («*/$) —r-60 -20 20 —r 80 100 120 -— T (°C)

Figure 2.5 The Prandtl number Pr, the kinematic viscosity

u and the

thermal conductivity A as a function of the temperature T

for:

air (A), water (B), methanol (C liquid, D vapour) and the

working pair lithium bromide zinc bromide - methanol (E)

In this the kinematic viscosity

v is

V = v / P [m

2

/s] (40)

The secondary passages of the exchangers in the AHP are completely fil led

with the liquid flow media, the primary passages, except for the mixture

heat exchanger, are not. In the absorber there is a (liquid) film flow over

the fins and a vapour flow in between. In the evaporator and condenser there

is a liquid and vapour flow, that is an evaporating and condensing film

flow. The generator is left out, this exchanger is not a CHME.

For heat exchangers operating with high-density fluids (in this case

liquids) the friction power is generally small relative to the heat transfer

rate, so has not a great influence. But for low-density fluids like air, as

one could see in the sections before, the friction power is of the same

order and even greater than the heat transfer rate. As said before, the heat

transfer coëfficiënt on the liquid side is 10 - 100 times greater than on

the vapour/gas side. So for the exchangers in the test plant, not the factor

j/f is of importance but the Colburn factor j.

The experimental result came always from a cross-flow configuration to keep

the header construction simple, in the exchangers of the AHP always

counter-flow took place. Only the position of the headers could give very a small

cross-flow influence.

In short, the thermodynamic properties and the phase of the chosen media as

well as the flow and corrugation configuration of the exchangers in the AHP

test plant form the major differences with the information found in

literature.

To find out whether or not the correlations found in literature for heat

transfer are in whatever form applicable for the heat transfer processes in

the exchangers of the test plant, a computer simulation model is developed

for a liquid flow over a offset strip fin surface.

In combination with the results of the experiments with the test plant heat

and mass transfer correlations will be searched which can be compared with

those found in literature. Both the experiments and the simulation model

will be discussed later.