Heat Exchanger Test Rig - Materials Technology · The development of a heat exchanger test rig capable of ... a project to the ... counterflow gas-gas heat exchangers. In this final

Heat ExchangerTest Rig

Erik van Kemenade

eindhoven university of technologydepartment of mechanical engineering

September 1999

This research was supported by the Netherlands Agency for Energy and the Environment(NOVEM)project 338420-7811

heat exchanger test rig i

Summary

The development of a heat exchanger test rig capable of measuring the effectivity of a counterflowgas-gas heat exchanger with an accuracy of 1 % is described.At an operating point defined by a secondary flow of 200 m3hr-1, balanced conditions and atemperature difference of 10 K between the ingoing temperatures of the primary and secondaryflows, the effectivity of a heat exchanger with a predicted effectivity of 89 % can be measuredwithin 1 %.

heat exchanger test rig i i

Samenvatting

De ontwikkeling van een test bank voor gas-gas tegenstroom warmtewisselaars waarmee hetmogelijk is om de effectiviteit met een nauwkeurigheid van 1 % te meten is beschreven.Op een werkpunt gedefinieerd door een secundaire stroming van 200 m3 hr-1, gebalanceerde conditiesen een temperatuur verschil van 10 K tussen de ingaande temperaturen, kan de effectiviteit van eenwarmtewisselaar met een effectiviteit van 89 % binnen 1 % worden gemeten.

heat exchanger test rig i i i

Contents

Summary i

Samenvatting i i

Contents iii

Symbols iv

1 Introduction 1

1.1 Counterflow gas-gas heat exchangers 11.2 Background of the project 21.4 Scope of the project 22 Effectivity measurement 3

2.1 Accuracy of the test rig 42.2 Temperature measurement 52.2.1 Thermistor properties 72.2.2 Design of the temperature measurement units 72.2.3 Flow conditioning 92.2.4 Velocity profile 112.2.5 Temperature profile and distance between the mixers 122.2.6 Influence of the velocity and the temperature difference 132.2.7 Number of mixers 132.2.7 Pressure drop 142.2 Mass- and capacity flow measurement 142.3 Ducts 162.4 Data acquisition 182.5 Predicted accuracy 183 Experimental validation 19

4 Conclusion and discussion 21

4.1 Conclusions 214.2 Discussion 21

Literature 22

heat exchanger test rig i v

Symbols

A heat exchanging surface [m2]C capacity flow [WK-1]D diameter [m]NTU number of transfer units [-]R H relative humidity [-]T temperature [K]U heat transfer coefficient [Wm-2K-1]

V volume flow [m3s-1]

e effectivity [-]e emissivityr specific mass [kgm-3]s standard deviation [-]

heat exchanger test rig 1

1 Introduction

In 1997 Novem commissioned a project to the Eindhoven University of Technology, Fontys/Ceditecand LEVEL energy technology to build a test rig for counterflow gas-gas heat exchangers. In thisfinal report the results are summarised.

1.1 Counterflow gas-gas heat exchangers

In recent years heat recovery techniques in industrial and domestic appliances have been gainingimportance due to the increased awareness of the limitations on the energy supply. Most heatrecovery techniques require a heat exchanger of some sort. A number of methods is used to recoverheat from exhaust flows such as ventilation air from buildings, damp hot air from dryers, or wastegases from burners. All these methods are designed to exploit the temperature difference betweenexhaust and supply flows to the full, using as little material and fan energy as possible. The mostcommon methods are recuperators such as cross-flow plate exchangers, regenerators such as heatwheels or alternating flow through porous masses and heat exchangers with an intermediatemedium, e.g. the twin coil or even the heat pump. These methods are still far from ideal, becausethey all allow mixing to take place so that the maximum temperature difference is not maintained.This drawback can be circumvented by using a counterflow recuperator.Counterflow recuperators have always been designed as plate exchangers. The heat transfer can bemuch improved by working with ducts instead of plates. Recently several patents have beenacquired concerning such a geometry, including one by a partner in this project, LEVEL energytechnology. Each duct is surrounded by a series of ducts in which the direction of flow is reversed(figure 1.1)

figure 1.1 counterflow duct recuperator

The recuperator is developed for a number of applications, including radiant burners, dryingprocesses and ventilation systems (Veltkamp 1995, van Kemenade 1997) with effectivities in theorder of 90 %, the effectivity being defined as the actual heat transfer divided by the maximumamount of heat which can be transferred.


1.2 Background of the project

During the (NOVEM supported) development of the LEVEL heat exchanger, it became evident thatit is very hard to measure the effectivity accurately. As the effectivity comes near the maximumvalue small deviations in the measured temperatures have a large impact on the recordedeffectivity. As the effectivity is directly related to the capacity of the heat exchanger defined bythe overall heat transfer coefficient U multiplied by the heat transferring surface A , it is animportant parameter to compare design alternatives and validate simulation tools. For that reasonNOVEM started a project to built a test rig enabling the determination of the effectivity with anaccuracy of 1 percent.

1.4 Scope of the project

The project is triggered by the development of high effectivity heat exchangers for ambientconditions (i.e. ventilation systems) and high temperature conditions (i.e. burners, gas turbines).The main goal is to develop techniques to determine the effectivity of such systems with an highaccuracy. The analysis tools developed during the project have a broad applicability, for practicalreasons the operating point of the test rig had to be limited however. Based on rather arbitraryconsiderations as availability of equipment and ease of construction, the operating point for the testrig is chosen at an air flow of 200 m3hr-1, balanced conditions and a temperature difference of 10 Kbetween the ingoing temperatures of the primary and secondary flows.The test rig is designed in an iterative process. First a parametric model is made for all the ingoingparameters which have an influence on the effectivity. Using this model (chapter 2) in combinationwith the experimental data for the components of the test rig, the accuracy's needed for thedifferent sensors were balanced to arrive at the desired accuracy. The experiments used aredescribed in chapter 3.In this report only the final results are given. Details can be found in the following reports:1998 Arink, C.A.F., Het ontwerp van een meetopstelling voor gas-gas warmtewisselaars,

Technische Universiteit Eindhoven, WOC-WET 98.002Contents: Development of the parametric model, initial choices for the measurementprinciples, validation of the bulk temperature measurement principle, functional design ofthe test rig.

1998 Huffel, M.L.P.C. van, De ontwikkeling van een warmtewisselaar meetbank, FontysHogescholen EindhovenContents: Error analysis for the data acquisition system, design of the data aquisition

1999 Kroonen, R.J.M.H., Het ontwerpen van een temperatuurmeetsectie om nauwkeurig detemperatuur van een luchtstroom in een kanaal te meten, Technische Universiteit Eindhoven,WOC-WET 98.002Contents: Development of the temperature measurement section, final design of the test rig

1999 Wolfs, M., De ontwikkeling van een warmtewisselaar meetbank, Fontys HogescholenEindhovenContents: Calibration of the individual sensors, development of the data acquisition system


2 Effectivity measurement

The effectivity of a heat exchanger is commonly defined as the actual heat transferred divided bythe maximum amount of heat which can be transferred. The maximum amount of heat which can betransferred is rather arbitrarily defined as the minimal capacity flow C (mass flow multiplied bythe heat capacity) times the maximum temperature difference available.

e =

Cc Tc,out - Tc,inCmin Th,in - Tk,in

=Ch Th,in - Th,outCmin Th,in - Tc,in

(2.1)

The subscript h denotes the hot fluid and c the cold. In principle the effectivity can be determinedfrom the bulk temperatures only as

e =

Ch Th,in - Th,outCh Thi,n - Th,in

=Th,in - Th,outThi,n - Th,in

when Ch < Cc (2.2)

e =

Cc Tc,out - Tc,inCc Thi,n - Tc,in

=Tc,out - Tc,inThi,n - Tc,in

when Cc < Ch (2.3)

e =

Th,in - Th,outTh,n - Tcin

=Tc,out - TcinThi,n - Tc,in

when Ch = Cc (2.4)

When the capacity flows of both fluids equal each other the heat exchanger is said to be balanced.During experiments this can be used as a reference case to check the equipment. When evaluatingequation (2.1) to (2.4) the heat capacity should be averaged over the temperature range of the heatexchanger. Normally however the physical properties are evaluated at the mean temperature.This condition is only satisfied if the heat exchanger is balanced as is shown in figure 2.1.

dA

surface

Cc < Ch

dTc

dTh

Th,in

Tc,out Th,out

Tc,in

DT

dA

surface

dTc

dTh

Th,in

Tc,out Th,out

Tc,in

DT

Cc > ChdA

surface

Cc = Ch

dTc dTh

Th,in

Tc,outTh,out

Tc,in

DT

figure 2.1 temperature profiles in a counterflow heat exchanger

The effectivity is closely related to the heat exchanging capacity of a heat exchanger defined asUA where U denotes the heat transfer coefficient and A the heat exchanging surface. The heatexchanging capacity is less dependant on the operating conditions than the effectivity. If the heatexchanging capacity is known for one operating point, the results can be used to calculate theperformance for other operating conditions. For a counterflow heat exchanger the relation betweenthe heat exchanging capacity and the effectivity is

e =

1 - exp 1 ±CminCmax

±UACmin

1 -CminCmax

exp 1 ±CminCmax

-UACmin

(2.5)

this equation reduces for a balanced counterflow heat exchanger to


e =

UACmin

1 +UACmin

(2.6)

These equation are plotted in figure 2.2

0 1 2 3 4 50

0.2

0.4

0.6

0.8

1Cmin/Cmax 0

0.250.5

0.751

NTU = UA/Cmin

effe

ctiv

ity

[-]

figure 2.2 effectivity of a counterflow heat exchanger as a function of the number of transfer units

2.1 Accuracy of the test rig

The accuracy of the determined effectivity is dependent on the accuracy in which the terms ofequation (2.1) can be determined. The accuracy's are given relatively, based on the relativestandard deviations sx/x derived of a series of measurements for x, except for the deviation in thetemperature which is given absolutely. All values given are based on the 95 % certainty intervalbounded by 2s, unless mentioned otherwise.

On the assumptions that the absolute standard deviations in the differential temperaturemeasurements sDT are constant and that the relative deviations in the capacity flows between thehot and the cold side are equal, the absolute error in the temperature and the relative error in thecapacity flows can be related to the accuracy in the effectivity for a chosen operating point (Arink1998, figure 2.3)

0

0.02

0.04

0.06

0.08

0.1

0 0.2 0.4 0.6 0.8 1relative error in the capacity flow [%]

abso

lute

err

or in

the

tem

per

atu

re d

iffe

ren

ce [K

]

± 0.5 %

± 1 %

± 1.5 %

± 2 %

figure 2.3 uncertainty in the effectivity as a function of the relative error in the capacity flow and the absolute errorin the temperature


Errors can be introduced in the piping system, peripheral equipment, and the measuring equipmentwhich are treated in Arink (1998). All possible sources for an inaccuracy in the measured effectivitymust be taken into account.The design of the test rig is divided in four parts- the heat exchanger is regarded as a black box of which the heat exchanging properties are

defined by the effectivity in one operating point. The heat exchanger is regarded as a "blackbox" with one major exception. It is inevitable that pressure losses occur when a medium flowsthrough the heat exchanger. The common description techniques for the performance treat heattransfer separately from the pressure drop. The pressure drop in a heat exchanger is irreversiblehowever and leads to an increase of the medium temperature. The measurements show thiseffect, while most analysis techniques do not take it into account. The solution can be found in asecond law analysis (Bejan 1985). To comply with current standards, in this report we willcorrect the measured temperatures with the measured pressure drop over the heat exchangerhowever. The effect has a comparatively large effect on the measured effectivity while theinfluence on the determined heat transfer coefficient is small.

- To determine the effectivity of a heat exchanger, temperatures, pressures, flows and humiditieshave to be measured.

- The ducts transport the air from the measurement stations to the heat exchanger. All heattransfer between the ducts and the surroundings and the pressure drop in the ducts influence themeasured temperatures.

- The data from the sensors have to be recorded in a suitable data acquisition systemAll subjects mentioned have an impact on the accuracy of the test rig. For instance, a less accuratetemperature measurement requires a more accurate pressure measurement or flow measurement. Notall iterations involved in this process are given in this report.

2.2 Temperature measurement

In order to determine the effectivity according to expression 2.1, the bulk temperatures of the airflow have to be measured on the locations indicated in figure 2.4. The problem to be solved is howthe measure the bulk temperature which is not identical to the mean temperature in a cross sectionas the velocities may differ. Consequently it is necessary to mix the airflow thoroughly beforemeasuring the temperature locally.Temperature sensors can be divided roughly in three categories as shown in figure 2.5. As air hardlyradiates optical temperature sensors are not suitable in this project. As the old mercurythermometer, still being used widely for calibration purposes, shows non-electrical thermometerscan be very good. Nevertheless electrical contact sensors are chosen as they allow for a relativelyeasy data acquisitionHaving made that decision, the option for a thermocouple and a resistivity method (plain or withNTC's) remains. Thermocouples, though trusty and widely applied, have a low temperaturedependence and a very low voltage output. It is very hard to get a accuracy below 1 K. Thedifference between the resistivity methods, represented by a Pt-100 sensor and two NTC thermistorsis shown in figure 2.5. The choice between a thermistor and Pt-100 was made rather subjectively onthe arguments listed in table 2.1. Thermistors are selected.


mixingsection

T mixingsection

T

RH

P

123456

V

heat exchanger

Tmixingsection

heater

mixingsection

T

RH

PV

123456

figure 2.4 schematic and sketch of the test rig

0.01

0.1

1

10

100

-20 0 20 40 60 80 100 120

RTD, Pt100

NTC, -3.0 %K-1NTC, -5.4 %K-1

temperature [oC]

rela

tive

res

ista

nce

RT/R

20o C

figure 2.5 relative resistance of a Pt100 and NTCsensor as a function of temperature

table 2.1 advantages and disadvantages ofNTC sensors compared with RTDsensors

Advantages Disadvantages

small dimensions non-linear

high resistance not standardised

high sensitivity lower temperature range

fast recalibration


2.2.1 Thermistor properties

Thermistors are intrinsic semi conductors produced from a ceramic material. The dependence of theresistance on the temperature is expressed as

R T = R T0 ×exp ±b 1T ±

1T0

(2.7)

NTC thermistors have a negative temperature dependence, the resistance globally varies from -3 to-6�% per Kelvin temperature rise. The value of the material constant b ranges from 500 to 2000 K.Most thermistors are produced as spheres or cylinders covered with a protective glass layer.Normally the thermistor are part of a wheatstone bridge, keeping either the current through- orthe voltage over the thermistor constant.As a current flows through the thermistor, ohmic heating will occur, which must be accounted for.Also the thermistor has to be connected and is subject to heat conduction through the wires. A thirderror may occur due to heat exchange with the surroundings.

2.2.2 Design of the temperature measurement units

The air flows are thoroughly mixed in the ducts (see ¤ 2.2.3) so that in principle one thermistor canbe used to measure the bulk temperature. Three thermistors are applied however to be able to detecta faulty thermistor. The design of the temperature measurement units are sketched in figure 2.6 andshown in figure 2.7 and 2.8 (The parameter variations given in this paragraph all start from thefinal design with a flow of 200 m3hr-1 and a temperature difference of 1 K between wall and airstream. The results are obtained from a model incorporating the following energy flows:- ohmic heating of the thermistor- convection to the thermistor- radiative heat transfer with the surroundings- conduction through the support- convection to the connection wires- ohmic heating of the connection wires- conduction through the connection wires.

figure 2.6 sketch of the temperature measurement unit


figure 2.7 temperature measurement unit figure 2.8 temperature sensor

The temperature of the thermistor depends on ohmic heating, the convective heat transfer from theair stream and the heat transfer with the surroundings due to radiation and conduction.The deviation of the thermistor temperature due to ohmic heating has a systematic nature and isbalanced by convective heat transfer . In figure 2.9 the deviation as a function of the air flow isgiven for the final construction.The thermistors are mounted within radiation shields consisting of a chromium plated brass tube(15x60 mm) to reduce the radiative heat exchange. The dimensions of the radiation shield are suchthat a large part of the heat transfer with the surroundings are suppressed with a minimumdisturbance of the convective heat transfer to the thermistor. Figure 2.10 gives a quantitativeimpression of the influence of the dimensions of the radiation shields on the readings from thethermistor when a temperature difference of 1 K exists between the duct walls and the air flow. Theradiation shields are mounted on PVC supports reducing the heat transfer to the wall. Due to thelow heat conduction of PVC, the dimensions of the support are not critical.

0

0.05

0.1

0.15

0.2

0.25

0.3

0.35

0.4

0 50 100 150 200 250

emissivity

1

0.70.1

tem

per

atu

re v

aria

tion

[mK

]

air flow [m3hr-1]

0

2

4

6

8

10

12

14

16

18

20

0 0.02 0.04 0.06 0.08 0.1

tem

per

atu

re v

aria

tion

[mK

]

length [m]

Dshield [mm]

5

15

25

figure 2.9 variation of the measured temperature as afunction of the flow and emissivity of theradiation shield

figure 2.10 variation of the measured temperature as afunction of the length and diameter of theradiation shield


The electrical connection of the thermistor plays a major role in the accuracy of the thermistor. Asheat- and electrical conductance are closely related, there exist no materials with a highelectrical- and a low thermal conduction. Thermal conduction can be reduced by applying a long thinwire but in that case the electrical resistance also rises which is undesired. Optimisation of botheffects resulted in an impractical small diameter or long length of the wire. For that reason anotherapproach is taken. A small wire ( 0.12 mm) is placed inside the air stream for a certain distancebefore it is firmly attached to the inner wall. At the inner wall it takes the wall temperature. For atemperature difference of 1 K between air and wall, the deviation of the thermistor temperaturedue to wire conduction is given in figure 2.11. A length of 0.22 m was chosen.The deviation of the thermistor temperature in the final design is plotted in figure 2.12 as afunction of the wall temperature and the flow, assuming a constant temperature of the air stream.

0

0.05

0.1

0.15

0.2

0.25

0 0.1 0.2 0.3 0.4 0.5

tem

per

atu

re v

aria

tion

[mK

]

wire length [m]

0

1

2

3

4

5

0 50 100 150 200 250

tem

per

atu

re v

aria

tion

[mK

]

air flow [m3hr-1]

DT [K]

15

5

1

figure 2.11 measured temperature as a function of thelength of the thermistor connection wires(D=0.12 mm)

figure 2.12 measured temperature as a function of theair flow and the temperature differencebetween the wall and the air stream

2.2.3 Flow conditioning

In the test rig, the secondary air temperature is heated to accomplish a temperature differencebetween the primary and secundary flow. The resulting air stream does not have a homogeneoustemperature profile however. For that reason the air stream is mixed to obtain a homogeneous(bulk) temperature. The desired mixing effectivity can be reached by applying turbulent mixers asare manufactured by Sulzer for instance (figure 2.13). Besides the Sulzer static mixer a home madevariety, designated by FT-150, was tested.


figure 2.13 static mixers, Sulzer (left) and FT-150 (right)

The experimental setup is shown in figure 2.14. The air stream stratified by heater comprising fourseparate convectors which can be regulated seperately. To make sure that a stratified flowdevelops, a flow straightener is placed behind the heater. The temperature distribution ismeasured before and after the mixing section with a ross containing 13 T-type thermocouples (figure2.15). In the mixing section several static mixers can be placed.

figure 2.14 experimental setup


figure 2.15 temperature measurement cross

Three temperature profiles represented by figure 2.16 were tested for all configurations.

1-0-0-0 1-1-0-0 0-1-1-0

figure 2.16 tempertaure profiles tested

2.2.4 Velocity profile

For one static mixer the axial velocities were measured behind the static mixers using a Pitot tube(figure 2.17 and 2.18). Immediately after the static mixer, the profiles are chaotic as can beexpected. After a settling length of one diameter (1D) the turbulent velocity profile starts todevelop and after 3D the velocity is almost constant. This distance is taken for the temperaturemeasurements. The measured deviations in the velocity profile are used to calculate the uncertaintyin the resulting bulk temperature.


●

●

●●

●

●

●

●

●

●●

●

●

●

●

●

❍ ❍ ❍

❍

❍❍ ❍

❍ ❍

❍ ❍ ❍❍

❍❍

❍

■

■ ■ ■■

■ ■ ■ ■ ■ ■ ■ ■■ ■

■▲

▲ ▲ ▲▲

▲▲ ▲ ▲

▲ ▲ ▲ ▲▲

▲▲

-1

0

1

2

3

4

5

6

7

0 20 40 60 80 100 120 140

● Dx = 0D❍ Dx = 1D■ Dx = 2D▲ Dx = 3D

position [mm]

velo

city

[ms-

1 ]

●

●

●

●

●

●

●●

●

●

●

●●

●

●

●

❍

❍

❍ ❍

❍

❍❍ ❍ ❍

❍ ❍❍

❍

❍

❍

❍

■

■

■■ ■ ■ ■

■ ■ ■ ■■ ■ ■ ■

■▲

▲▲ ▲ ▲ ▲

▲▲ ▲ ▲

▲▲ ▲ ▲ ▲

▲

-1

0

1

2

3

4

5

6

7

0 20 40 60 80 100 120 140

● Dx = 0D❍ Dx = 1D■ Dx = 2D▲ Dx = 3D

position [mm]

velo

city

[ms-

1 ]

figure 2.17 axial velocity profile after the Sulzer mixer figure 2.18 axial velocity profile after the FT-150mixer

2.2.5 Temperature profile and distance between the mixers

In the next experiment two mixers were used at several distances between each other and for thetemperature profiles of figure 2.16. The Sulzer mixer is not very sensitive to variations in the initialtemperature profile, the mixing effectivity is between 95 and 98 % in all cases. The FT-150 mixer ismore sensitive to the inlet temperature profile (figure 2.19 and 2.20). Subsequently the 1-1-0-0profile was used as a worst case situation. For both mixer types the best mixing effectivity isreached when the air stream is allowed to settle for 1D between the two mixers. If the distance isless, the chaotic temperature profile hinders the second mixer, if the distance is larger,stratification occurs.

●

●

● ● ●● ●

❍❍ ❍

❍

❍

❍ ❍

■

■■

■■ ■

■

92

93

94

95

96

97

98

99

100

0 0.5 1 1.5 2 2.5 3

mix

ing

effe

ctiv

ity

[-]

distance between mixerspipe diameter

1-0-0-0

1-1-0-0

0-1-1-0

●

●

●

●

●●

●

❍

❍❍

❍❍

❍ ❍

■

■

■■

■

■ ■

92

93

94

95

96

97

98

99

100

0 0.5 1 1.5 2 2.5 3

mix

ing

effe

ctiv

ity

[-]

distance between mixerspipe diameter

1-0-0-0

1-1-0-0

0-1-1-0

figure 2.19 mixing effectivity after two Sulzer Mixersas a function of the distance between themixers for the temperature profiles offigure 2.16

figure 2.20 axial velocity profile after the FT-150mixer as a function of the distance betweenthe mixers for the temperature profiles offigure 2.16


2.2.6 Influence of the velocity and the temperature difference

In figure 2.21 and 2.22 the influence of the velocity on the mixing effectivity is shown. The velocityhas almost no influence on the mixing effectivity. The influence of the initial temperaturedifference is also small.

● ●●

●

❍ ❍

❍

❍

■

■

■

■

▲▲

▲

▲

93

93.5

94

94.5

95

50 100 150 200 250flow [m3hr-1]

mix

ing

effe

ctiv

ity

[-]

DT [oC]

4

9.5

2015.5

●

● ●●

❍

❍

❍❍

■

■

■■

▲

▲

▲

▲

93

93.5

94

94.5

95

50 100 150 200 250flow [m3hr-1]

mix

ing

effe

ctiv

ity

[-]

DT [oC]

9.5

20

15.5

4

figure 2.21 mixing effectivity after two Sulzer mixerswith a spacing of 1D as a function of the airflow and temperature difference

figure 2.22 mixing effectivity after two FT-150 mixerswith a spacing of 1D as a function of the airflow and temperature difference

2.2.7 Number of mixers

The number of mixers in combination with their spacing determines the mixing effectivityachieved. This influence is tested within the rather arbitrary constraint that the length of themixing section should be below 8D. For this experiment only FT-150 mixers were used. Though theSulzer mixer arguably has a better performance, the FT-150 mixer also performs adequately and isreadily available. The configurations of figure 2.23 were tested and configuration 7 is chosen as ithas a good mixing effectivity and a low pressure drop.

X-D-X-5.4D

X-D-X-D-X-3.6D

X-X-D-X-4.6D

X-X-X-O-X-4.6D

X-O-X-O-X-O-X-1.8D

X-X-O-X-X-3.8D

X-X-O-O-X-X-2.8D

X-X-X-X-X-4D

X-X-X-X-X-4D

X-X-X-O-X-X-3D

X-X-X-X-X-X-3.2D

0 20 40 60 80 100 120 90 92 94 96 98 100

configuration

pressure drop [Pa] mixing effectivity [%]

figure 2.23 pressure drop and mixing effectivity for the tested mixer configurations


2.2.7 Pressure drop

Using the static mixers leads to an irreversible pressure loss and heat generation. This is no problemat the inlets of the heat exchanger but the temperature at measured at the outlet has to be correctedfor this heat generation. The increase of the temperature is calculated according to the pressure-flow relation of figure 2.24

0

100

200

300

400

500

0 50 100 150 200 250

●

●

●

●

●

●

flow [m3hr-1]

pre

ssu

re d

rop

[Pa]

figure 2.24 pressure drop over the mixing section as a function of the flow

2.2 Mass- and capacity flow measurement

The mass flow can be measured directly (hot wire, Corioli) but these principles cannot meet thedemanded accuracy. Volume flow measurement can be done by positive displacement meters andtechniques where in a cross section the time and position averaged velocity is measured. Positivedisplacement meters can reach the best accuracy. An IGA rotary displacement flow meter is used(figure 2.25). Measurement errors are mainly due to internal leak as deviations in the countermechanism can be calibrated.

figure 2.25 principle of a rotary displacement flow meter

To convert the measured volume flow to a mass flow , the pressure, temperature and relativehumidity must be known (and can introduce a deviation) to assess the correct values for the specificmass and heat capacity of the air flow. In the test rig, the temperature is measured directly beforethe gas meter with a temperature measurement unit as described in ¤ 2.2.2.


The gas meter has a rather large heat capacity so care must be taken that the meter is at the correcttemperature. Neglecting heat transfer with the surroundings, the heat balance for the gas meterbecomes

kA Tf±Tbulk =Mfcp f

dTfdt

(2.8)

kA denotes the heat exchanger ÒsizeÓ, and Mf the mass of the heat exchanger.From this equation the time constant t can be calculated as

TbulkTf

t = 1t×t+1 t =

Mfcp fkA

(2.9)

indicating that after t seconds, about 40 % of the initial temperature difference remains. In Arink(1998) the time constant is calculated at half an hour. To prevent deviations in the temperaturemeasurement sections due to radiative exchange with the flow meter, it is advised to wait for twohours before commencing with the measurements.To convert the measured volume flow rate to a mass rate, the specific mass and heat capacity haveto be determined at the pressure, humidity and temperature near the gas meter. The temperature isknown from the temperature measurement units in the outlets. The pressure is determined using thelaboratory mercury barometer in combination with calibrated Huba control differential pressuregauges. In Wolfs(1999) is shown that the relative deviation of the pressure gauge remains within0.1 %.A capacitative humidity sensor is applied. The sensor was tested by comparing it with a (wet bulb)thermometer (Wolfs 1999). The relative deviation can amount to 2%. Humidity sensors arenotorious for errors due to saturation or fouling. During start-up of the test rig a simple check can beperformed by comparing the readout of the R.H. sensor to the theoretical values. In figure 2.26 and2.27 the readings of the temperature and humidity sensors is given during start-up. Theoreticallythe relative humidity should drop from an initial value of 0.023 to 0.008 which indeed is indicatedby the R.H.. sensor.

0 1 2 3 4 520

24

28

32

36

40

tem

per

atu

re [°

C]

time [103 s]0 1 2 3 4 5

0

0.05

0.1

0.15

0.2

0.25

rela

tive

hu

mid

ity

[-]

time [103 s]

figure 2.26 mean temperature as a function of the timeduring start-up

figure 2.27 relative humidity as a function of the timeduring start-up


2.3 Ducts

While flowing through the ducts the air will change in temperature due to heat transfer with thesurroundings. These effects have been modelled by both Arink (1998) and Kroonen (1999). Concerningthe heat transfer, the forced convection on the inside of the pipe is modelled using the well knownChurchill-Chu empirical relation for the Nusselt number, for the natural convection on the outsidethe Seider-Tate equation is applied. The radiative heat transfer on the inside is calculated bydividing the pipe in segments, assuming that axial heat transfer through wall conduction isnegligible compared to the radial conduction. In figure 2.28 to 2.32 the influence of several designvariables are given, starting from the conditions of table 2.2.

table 2.2 reference conditions

geometry, material properties operating conditions

piping insulation

inner diameter 0.15 m thickness 0.05 m flow 225 m3hr-1

wall thickness 0.005 m conduction coef. 0.038 Wm-1K-1 temperature 20 °C

length 5 m emission coefficient 0.1 inlet temperature 20 °C

0

0.05

0.1

0.15

0.2

0.25

0.3

0 0.01 0.02 0.03 0.04 0.05 0.06 0.07

heat conduction coefficient [Wm-1K-1]

tem

per

atu

re v

aria

tion

[Km

-1]

0

0.05

0.1

0.15

0.2

0.25

0.3

0.35

0 0.05 0.1 0.15

tem

per

atu

re v

aria

tion

[Km

-1]

insulation thickness [m]

figure 2.28 influence of the insulation heat conductioncoefficient on the axial temperaturevariation

figure 2.29 influence of the insulation thicknesscoefficient on the axial temperaturevariation


0

0.1

0.2

0.3

0.4

0.5

0 0.2 0.4 0.6 0.8 1

tem

per

atu

re v

aria

tion

[Km

-1]

emission coefficient [-]

0

0.5

1

1.5

2

2.5

0 50 100 150 200 250 300air flow [m3hr-1]

tem

per

atu

re v

aria

tion

[Km

-1]

figure 2.30 influence of the insulation emissioncoefficient on the axial temperaturevariation

figure 2.31 influence of the air flow on the axialtemperature variation

-0.1

-0.05

0

0.05

0.1

0.15

0.2

0.25

0.3

-10 -5 0 5 10 15 20 25 30

tem

per

atu

re v

aria

tion

[Km

-1]

temperature [oC]

figure 2.32 influence of the air inlet temperature on the axial temperature variation

The material and thickness of the insulation layer used have a large influence on the airtemperature in the ducts. For instance: a decrease in the heat conduction coefficient from 0.05 to0.025 Wm-1K-1 results in a 40 % lower temperature drop. The effect of enlarging the thickness of theinsulation is limited. Increasing the thickness also leads to a larger surface and there exists anoptimum layer thickness. The thickness of the duct itself and the radiative properties of its innersurface are hardly relevant. The outside emissivity e is of importance as radiative heat transfercan account for 18 % (e=0.1) to 69% (e=1) of the total heat transfer.Based on these considerations PVC is chosen as a suitable and readily available material for theducts. The selection of the insulation material is based on the heat conduction coefficient anddiffusion resistance to prevent condensation. Extruded polystyrene with a closed cell structure hasgood diffusion resistant properties (m = 7500) and a low heat conduction coefficient (l = 0.038Wm-1K-1). Radiation can be reduced by adding a reflective layer of aluminium foil.


2.4 Data acquisition

The data acquisition system is documented in Wolff (1998) and van Huffel (1998). As there is noneed for a fast measurement system, a National Instruments data acquisition system based on SCXImodules (signal condition extensions for instruments) is used. The acquisition part of modules for thetemperature, pressure, volume flow and pressure measurements.The pressure- and relative humidity measurements are straightforward, the sensors deliver asignal in the range between 0-10 V and the accuracy can be calculated from the accuracy of thecomponents (power supply, sensor and measuring device).The resistances of the RTD's are measured using a wheatstone bridge. The wheatstone bridge has tobe matched to provide the conditions mentioned in paragraph 2.2 and to provide an optimal signalfor the data acquisition system. Deviations which may occur in the resistances used add to the totalaccuracy of the system.The flow meters deliver a pulse signal using a custom build segment disc and optical sensor (vanHuffel 1998). During each measuring cycle the number of pulses is counted during a specified timeinterval. The accuracy of this procedure is determined by the length of the counting interval. Thistime interval is chosen such that the resulting error is negligible compared with the deviations inthe flow meter itself.To facilitate monitoring of the measurement process, the effectivity and mass flows are calculatedon line, while the raw data are stored. The final results are obtained using a separate post-processing program. This procedure allows backtracking of the results whenever a measurementanomaly is observed.

2.5 Predicted accuracy

The deviations which may occur due to the effects quantified in the preceding paragraphs arecombined in a numerical model to predict the accuracy of the test rig for a specified operating point(Kroonen 1999). The predicted error assuming that systematic errors are avoided is 0.82 %, with acertainty of 95 %. The contribution of some main parameters to this error is depicted in figure 2.33.

0 0.05 0.1 0.15 0.2 0.25 0.3

massflow

massflow

heat capacity

heat capacity

temperature

temperature

prim

ary

flow

secu

ndar

yfl

ow

relative contribution to the total random deviation

figure 2.33 contribution of the major parameters to the error in the accuracy


3 Experimental validation

The experiments performed to validate the test rig are given in table 3.1. A resume of the mainresults is given in this chapter. For the measurements a LEVEL counterflow heat exchanger wasused.

table 3.1 experiments

experiment motivation method

flow meter calibration elimination of systematic errors,checking random errors

comparison with a reference floemeter (Wolfs 1999)

pressure difference gaugecalibration

elimination of systematic errors,checking random errors

comparison with a Betz micro-manometer (Wolfs 1999)

RH sensor check check if the error remains within2�%

comparison with a wet bulb RHsensor (Wolfs 1999)

Temperature sensor

1 power supply stability/noise check monitoring with an oscilloscope(Kroonen 1999)

2 wiring noise check monitoring with an oscilloscope(Kroonen 1999)

3 calibration determining the resistance-temperature relation

EUT calibration service (Kroonen1999)

4 comparison betweenmeasurement sections

check if the assumptions for themeasurement unit hold

comparison between themeasurement sections (Kroonen1999)

Heat exchanger

1 effectivity deviation < 1 % check the theory bring the heat exchanger in balancein check the effectivities (Kroonen1999)

2 working range check the accuracy of the test rigoutside the operating point

repeat the measurements for otherflows

All the influences mentioned in the preceding sections are modelled to predict the deviations in thetemperature for the test rig. The model has been validated by placing two identical measurementsections behind each other. The experimental results are given in figure (3.1) The thermistors werecalibrated against each other at a flow of 200 m3hr-1 and a temperature of 20 *C. For thesecondition the temperature rise in the measurement section is as expected. At lower flows thepredicted temperature difference is smaller and vice versa at higher flow rates. This artefact isprobably due to the calibration procedure.


●

●

●

●●

●

❍

❍ ❍

❍

❍

■

■

■

-0.15

-0.1

-0.05

0

0.05

0.1

0.15

-0.2 -0.1 0 0.1 0.2DTpredicted [°C]

DT

mea

sure

d [°

C]

flow [m3hr-1]●

■

❍

141180207

temperature [°C]

29.7

25.5

22.020.2

19.218.9

29.8

27.4 23.420.7

18.1

24.4

22.0

18.2

figure 3.1 comparison between the predicted and measured temperatures

The final test was a comparison with a LEVEL laminar counterflow recuperator. Both theeffectivity and the pressure loss was measured (figure 3.2 and 3.3). The measured effectivitymatches the calculated values at the point where the thermistors were calibrated. At lower valuesthe measured effectivity starts to deviate as is to be expected. The pressure drop is somewhat largerthan calculated and shows a turbulent profile combined with a turbulent profile. This is due to thepressure losses associated with the in- and outflow ports of the heat exchanger casing.

✖

✖

✖

✖✖

0.8

0.85

0.9

0.95

1

0 50 100 150 200 250

calculated

measured

effe

ctiv

ity

[-]

flow [m3hr-1]

✖

✖✖

✖

✖✖

0

20

40

60

80

100

120

140

0 50 100 150 200 250

✖

calculated (LEVEL)A

in- and exit lossesB

A+Bmeasurements

flow [m3hr-1]

effe

ctiv

ity

[-]

figure 3.2 calculated and measured effectivity asfunction of the flow

figure 3.3 calculated and measured pressure drop asfunction of the flow


4 Conclusion and discussion

4.1 Conclusions

Ñ At an operating point defined by a secondary flow of 200 m3hr-1, balanced conditions and atemperature difference of 10 K between the ingoing temperatures of the primary andsecondary flows, the effectivity of a heat exchanger with a predicted effectivity of 89 % canbe measured within 1 %.

Ñ The numerical model predicting the accuracy of the test rig can be used to assess the maximumeffectivity in other operation points. In general the accuracy is higher when the (measured)effectivity is lower, either by changing the heat exchanger design or changing the operatingconditions.

Ñ The initial assumption that the heat exchanging power UA of a heat exchanger can bedetermined by measuring the temperatures and the flows is false. The irreversible lossesassociated with the pressure drop and subsequent temperature rise of the medium is not takeninto account in the definition of the effectivity.

4.2 Discussion

It is proven that it is possible to measure the effectivity of a heat exchanger with an accuracy of1�%, but only within strictly specified conditions. Consequently, quoting an effectivity withoutspecifying the conditions is rather useless.As is, the test rig is only proven to determine the effectivity within one percent at one operatingpoint and for a certain heat exchanger. The next step will be to asses the accuracy of the test rig inother operating conditions and for heat exchangers with a different heat exchanging capacity.The operating range of the test rig is limited by the sensors and materials used. The analysis of thedeviations which occur in the test rig show that it is hardly feasible to construct a test rig which iscapable to determine the effectivity of a broad range of heat exchangers in a broad field ofoperating conditions. Also is shown however that it is possible to predict the accuracy achievedunder certain conditions. This allows for a fast design of a test rig which is applicable for thedesired conditions.During the project it became evident that the generally adopted definition of the effectivity islacking because irreversible pressure losses are neglected as is recognised by several authors (Soland1978, Bejan 1988, Veltkamp 1993). The solution adopted during this project is to correct the measuredtemperatures with the measured pressure losses to comply with the current practice. Anothersolution can be to adopt a characterisation of the heat exchanger which includes the pressure lossesusing a second law analysis as is proposed by Bejan (1988).


Literature

1978 Soland, J.G., W.M. Mack and W.M. Roshenow, Performance ranking of plate-fin heatexchanging surfaces, J. Heat transfer vol. 100, p. 514

1988 Bejan, A, Advanced engineering thermodynamcs, McGraw-Hill, New York1993 Veltkamp, W.B., Haalbaarheidsstudie van een laminaire tegenstroom kanaal

warmtewisselaar, LEVEL enrgietechniek rapport 93.09, Son1997 Kemenade, E van, Ontwikkeling prototype recuperatieve stralingsbrander , LEVEL

enrgietechniek rapport 97.01, Son1998 Arink, C.A.F., Het ontwerp van een meetopstelling voor gas-gas warmtewisselaars,

Technische Universiteit Eindhoven, WOC-WET 98.0021998 Huffel, M.L.P.C. van, De ontwikkeling van een warmtewisselaar meetbank, Fontys

Hogescholen Eindhoven1999 Kroonen, R.J.M.H., Het ontwerpen van een temperatuurmeetsectie om nauwkeurig de

temperatuur van een luchtstroom in een kanaal te meten, Technische Universiteit Eindhoven,WOC-WET 98.002

1999 Wolfs, M., De ontwikkeling van een warmtewisselaar meetbank, Fontys HogescholenEindhoven

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Heat Exchanger Test Rig - Materials Technology · The development of a heat exchanger test rig capable of ... a project to the ... counterflow gas-gas heat exchangers. In this final