NUREG/IA-0184

International 

Agreement Report

In-Tube Steam Condensation 

in the Presence of Air 

Prepared by 

A. Tanrikut/TAEA 

0. Yesin/METU 

Turkish Atomic Energy Authority 

Eskisehir Yolu 06530 Ankara, Turkey 

Middle East Technical University 

06531 Ankara, Turkey 

Office of Nuclear Regulatory Research 

U.S. Nuclear Regulatory Commission 

Washington, DC 20555-0001 

June 2000 

Prepared as part of 

The Agreement on Research Participation and Technical Exchange 

under the International Code Application and Maintenance Program (CAMP) 

Published by 

U.S. Nuclear Regulatory Commission
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NUREG/IA-0184

International 

Agreement Report

In-Tube Steam Condensation 

in the Presence of Air 

Prepared by 

A. Tanrikut/TAEA 

0. Yesin/METU 

Turkish Atomic Energy Authority 

Eskisehir Yolu 06530 Ankara, Turkey 

Middle East Technical University 

06531 Ankara, Turkey 

Office of Nuclear Regulatory Research 

U.S. Nuclear Regulatory Commission 

Washington, DC 20555-0001 

June 2000 

Prepared as part of 

The Agreement on Research Participation and Technical Exchange 

under the International Code Application and Maintenance Program (CAMP)

Published by 

U.S. Nuclear Regulatory Commission
ABSTRACT

In this research work, in-tube condensation in the presence of air is investigated 

experimentally for different operating conditions, and inhibiting effect of air is analyzed by 

comparing the experimental data of air/steam mixture with the data of corresponding pure 

steam cases, with respect to temperature, heat flux, air mass fraction, and film Reynolds 

number. The test matrix covers the range of; Pn=2-6 bars, Re,=45,000-94,000, and 

Xi=0 % - 52 %. The inhibiting effect of air manifests itself as a remarkable decrease in 

centerline temperature (10 'C - 50 °C), depending on inlet air mass fraction. However, the 

measured centerline temperature is suppressed compared to the predicted one, from the Gibbs

Dalton law, which indicates that the centerline temperature measurements are highly affected 

by inner wall thermal conditions, possibly due to narrow channel and high vapor Reynolds 

number. Even at the lowest air quality (10 %) the reduction of the local heat flux is 20 % 

while it reaches up to 50 % for the quality of 40 %. Vapor mass flow rate may dominate over 

system pressure, concerning the effect on local heat flux, for cases with air/vapor mixture. The 

situation is rather different in pure vapor runs, that is increase in system pressure has a strong 

effect on enhancement of predicted, and even measured, wall subcooling degree and hence on 

increase in local heat flux. The investigation for the effect of superheating of steam on 

condensation process reveals the fact that inlet superheating of steam has no considerable 

effect on heat flux. The calculated film Reynolds number decreases as air mass fraction 

increases at the same system pressure setting, and falls into the range of turbulent region 

(Ref > 300) for almost all experimental runs. The RELAP5 code overpredicted majority of 

experimental local heat flux data by 5 % - 50 %.

iii
TABLE OF CONTENTS

ABSTRACT iii 

TABLE OF CONTENTS v 

NOMENCLATURE vii 

CHAPTER 

1. INTRODUCTION 1 

2. DESCRIPTION OF THE TEST FACILITY 3 

2.1 Steam/gas Supply 3 

2.2 Connecting Piping and Pipe Fittings 4 

2.3 Test Section 5 

2.3.1 Condenser Tube 5 

2.3.2 Jacket Pipe 6 

2.4 Instrumentation 7 

2.4.1 Thermocouples 7 

2.4.2 Pressure Transducer 12 

2.4.3 Flowmeter 13 

2.5 Data Acquisition System 14 

3. OPERATING PROCEDURES OF THE TEST FACILITY 15 

3.1 System Check 15 

3.1.1 Isothermal Check of Thermocouples 15 

3.1.2 Prediction of Environmental Heat Loss 17 

3.1.3 Prediction of Fin Effect for Inner Wall Temperature 

Measurements 17 

3.1.4 Reproducibility of Data 20 

3.2 Experiments 21 

3.2.1 System Start-up 21 

3.2.2 Operating at Steady-state Conditions and 

Data Logging 22 

3.2.3 System Shutdown 22 

4. EXPERIMENTAL TEST MATRIX 23 

5. DATA REDUCTION PROCEDURE 26

V
6. EXPERIMENTAL RESULTS AND DISCUSSION 30 

6.1 Introduction 30 

6.2 Temperature Distribution 30 

6.3 Local Heat Flux Distribution 37 

6.4 Local Air Mass Fraction Distribution 44 

6.5 Condensate Film Reynolds Number 46 

6.6. Comparison with Theory 47 

LIST OF REFERENCES 50 

APPENDIX A: PHOTOGRAPHS OF THE METU-CTF 51 

APPENDIX B: SPECIFICATIONS OF INSTRUMENTATION 

AND DATA ACQUISITION SYSTEMS 55 

APPENDIX C: ERROR ANALYSIS 56

vi
NOMENCLATURE

Latin Symbols: 

A Coefficient of exponential fitting expression 

A Area, m2 

B Coefficient of exponential fitting expression 

Cp Specific heat at constant pressure, J/kg-°C 

d, D Diameter, m 

h Convective heat transfer coefficient, W/m 2

-°C 

h Enthalpy, J/kg 

rh Mass flow rate, kg/s 

M Molecular weight 

P Pressure, bar 

q Heat transfer rate, W 

q" Heat flux, W/m2 

r Correlation coefficient 

Re Reynolds number 

S Deviation from fitting curve 

T Temperature, 'C 

x Axial distance, m 

X Noncondensable gas quality 

Greek Symbols: 

(X Flow coefficient 

Ax Dynamic viscosity, kg/m-s 

p Density, kg/M3 

( Standard deviation 

Subscripts: 

cw Cooling water 

D Based on diameter 

h Hydraulic 

i Inner, inlet 

n Nominal 

s Saturation 

t Total 

v Vapor phase 

w Wall

vii
CHAPTER 1

INTRODUCTION 

The introduction of nuclear power becomes an attractive solution to the problem of 

increasing demand for electricity power capacity in Turkey. Thus, Turkey is willing to follow 

the technological development trends in advanced reactor systems. A part of our long-term 

research and development efforts is planned to concentrate on passive cooling systems. The 

primary objectives of the passive design features are to simplify the design, which assures the 

minimized demand on operator, and to improve plant safety. The research on passive systems 

mainly comprises the computer code assessment studies and includes the applications for both 

old and new generation reactor systems. To accomplish these features the operating principles 

of passive safety systems should be well understood by an experimental validation program. 

Such a validation program is also important for the assessment of advanced computer codes, 

which are currently used for design and licensing procedures. The condensation mode of heat 

transfer plays an important role for the passive heat removal applications in the current 

nuclear power plants (e.g. decay heat removal via steam generators in case of loss of heat 

removal system) and advanced water-cooled reactor systems. But it is well established that the 

presence of noncondensable gases can greatly inhibit the condensation process due to the 

build-up of noncondensable gas concentration at the liquid/gas interface. The isolation 

condenser of passive containment cooling system of the simplified boiling water reactors is a 

typical application area of in-tube condensation in the presence of noncondensable. The 

research work concerning the application of condensation in the presence of air, as a 

noncondensable gas, was first undertaken for a Once Through Steam Generator (OTSG) type 

of PWR for which experimental data were available. These experimental data were obtained 

from the 2X4 test loop of the University of Maryland at College Park (UMCP), addressing a 

very important safety issue called the loss of residual heat removal system after reactor 

shutdown. The experimental data were used for the assessment of RELAP5/mod3 (v5m5) 

thermal-hydraulic computer code and both the effect of Nusselt model, incorporated in the 

code as the condensation model, and the effect of nodalization model were investigated [1].

1
But the lack of measurements for the inside of the steam generator has led us to the conclusion 

that the separate effect test is strongly needed for the investigation of in-tube condensation and 

the effect of noncondensables on the condensation mode of heat transfer. Thus, an 

experimental study which could enable us for the fundamental investigation of condensation 

in the presence of air was planned in cooperation with the Mechanical Engineering 

Department of the Middle East Technical University (METU), Ankara, in the frame of a 

project (Project No: 94403507) between the Turkish Atomic Energy Authority (TAEA) and 

METU. The project is partially sponsored by the International Atomic Energy Authority 

(IAEA) under the Coordinated Research Program (Contract No: 8905/RO) which is entitled 

"Thermohydraulic Relationships for Advanced Water Cooled Reactors '" 

The experimental program [2] covers a wide range of steam and air/steam mixture 

flow rates under forced convection conditions -which partly falls outside the range 

encountered in typical passive heat removal applications in NPPs due to high Reynolds 

number- and has the purpose to investigate the inhibiting effect of air on steam condensation 

process. The results of this experimental study are also planned to be supplementary in nature 

for other experimental investigations such as those performed at the Massachusetts Institute of 

Technology (MIT), Cambridge, and University of California, Berkeley, (UCB). These 

investigations undertaken at MIT and UCB aim to support GE's Passive Containment Cooling 

System (PCCS) and Isolation Condenser (IC) designs, with relatively lower Reynolds number. 

As stated in the CAMP (Thermalhydraulic Code Applications and Maintenance 

Program) agreement between the Turkish Atomic Energy Authority (TAEA) and the US 

Nuclear Regulatory Commission (USNRC), the experimental data on in-tube steam 

condensation in the presence of air are opened to the USNRC. The data will be utilized for the 

assessment and validation of the RELAP5/mod3 thermal-hydraulic system analysis code. The 

assessment of this computer code, based on the data of MIT and UCB for pure steam 

condensation and condensation in the presence of air and helium was finished. [3].

2
CHAPTER 2

DESCRIPTION OF THE TEST FACILITY 

The test facility, named as METU Condensation Test Facility (METU-CTF), was 

installed at the Mechanical Engineering Department of the Middle East Technical University 

(METU). The photographs of the facility are presented in Appendix A. The experimental 

apparatus consisting of an open steam or steam/gas system and an open cooling water system 

is depicted in the flow diagram of Figure 2.1. The details of the apparatus are described in the 

following Sections: 

2.1 Steam/gas Supply 

Steam is generated in a boiler (1.6 m high, 0.45 m ID) by using four immersion type 

sheathed electrical heaters. Three of these heaters have a nominal power of 10 kW each and 

the fourth one has a power of 7.5 kW, at 380 V. All the heaters can be individually controlled 

by switching on or off. One of these heaters, i.e. the one with 7.5 kW power, is connected to a 

variac for continuous control of power. 

The boiler tank was designed to withstand an internal pressure of 15 atm (at T=20 °C) 

and was tested at this pressure. The maximum operating pressure of the tank is 10 atm. To 

ensure dry steam at the exit of the boiler, a mechanical separator directly connected to the exit 

nozzle was installed. However, electrical pre-heating with three heaters (0.5 kW per heater) is 

also available at the entrance of the test section to increase the temperature of steam, so that 

steam is guaranteed to be 100% dry. The boiler tank was thermally insulated to reduce 

environmental heat loss. 

Compressed air can be supplied either to the boiler tank (directly to the water) or to the 

steam line via a nozzle (after the orifice meter) on the horizontal part of the pipe which 

connects the boiler and the test section. Preference was given to the first method; i.e. injection 

to the boiler, during most of the experiments since system behavior is more stable compared

3
to the second method, when air mass flow rate is increased. When air injection was performed 

by the second method (to the horizontal piping), air injected passes through the preheating 

section so that local steam condensation was avoided at the entrance of the test section due to 

thermal inequilibrium of steam and air. The air supply system consists of an air compressor 

and three compressed air tanks with a total capacity of 600 liters. The maximum pressure of 

the compressed air system is 10 bars. 

The boiler tank is equipped with the measuring instruments given below; 

- level gauge with an operating pressure of 16 bars and a test pressure of 32 bars, 

- safety vent valve of spring lift type with an operating pressure of 12 bars, 

- pressure controller for cutting the power off at a predetermined maximum pressure 

setting, 

- pressure gauge (1-16 bars), 

- relief valve (19.05 mm ID). 

2.2 Connecting Piping and Pipe Fittings 

The pipe connecting the boiler tank and the test section has a length of approximately 

2 m and an ID of 38.1 mm. The pipe was connected to the boiler tank via an isolation valve. 

This isolation valve (38.1 mm ID) is used to isolate the boiler until inside pressure of the tank 

is increased to a pre-determined level. The measurements performed on this part of the 

experimental facility are mass flow rate via a differential pressure transmitter and temperature. 

There are three electric heaters (0.5 kW each at 220 V) installed to the horizontal part of the 

piping between the orifice meter and the test section. The pipe connecting the boiler and the 

test section was thermally insulated.

4
ID: 38. Imm 

L" 2 6 00nu

Jackct Pipe 

(8 1/89 ID/OD, L: 2133 mm)

makeup %vx Uir ne (D T*W l Prb 

To Drain (® P Tre T.du zns• 

Figure 2. 1. The Flow Diagram of the METU-CTF 

2.3 Test Section 

The test section is a heat exchanger of countercurrent type, that is steam or steam/gas 

mixture flows downward inside the condenser tube (inner tube) and cooling water flows 

upward inside the jacket pipe (outer pipe). 

2.3.1 Condenser Tube 

The condenser tube consists of a 2.15 m long seamless stainless steel tube with 

33/39 mm ID/OD and is flanged at both ends with sealing materials. The condenser tube was 

flanged to the inlet (33.5/42.6 mm ID/OD) and exit (33.5/42.6 mm ID/OD) pipes of the test

5
section. The total length of the inlet pipe from the horizontal part of the pipe section down to 

the condenser tube is approximately 33 cm (10 x di , where di is the inner diameter of the tube) 

and this length is long enough for the mixture flow to become fully developed before entering 

the condenser. It should also be noted that some uncertainties (such as irregular film 

development or dropwise condensation) associated with the liquid film development at the 

entrance of the condenser tube are expected to occur in this development region since the 

entrance region was not thermally insulated. A pressure measurement port was located at the 

vertical part of the inlet pipe flanged to the condenser tube. 

A total of 13 holes (1.5 mm diameter) were drilled with an angle of 300 at different 

elevations along the condenser tube length to fix the thermocouples for inner wall temperature 

measurements. The condenser tube was tested at 10 atm pressure to check that inner wall of 

the tube was not pierced during the drilling process. 

The outlet of the condenser tube is connected to a tank via exit part of the test section. 

This tank is used to keep the system pressure at a constant level by controlling the flow rate of 

steam or air/steam mixture through a valve connected to the tank. The measured parameters at 

the exit of the test section are pressure and temperature. 

2.3.2 Jacket Pipe 

The jacket pipe surrounding the condenser tube is made of sheet iron and has a length 

of 2.133 m and 81.2/89 mm ID/OD. The cooling water is supplied via a nozzle which has 

been welded on the jacket pipe. Similarly, cooling water outlet consists of a nozzle which is 

connected to the building water discharge system. Inner diameter of all these nozzles is 12.7 

mm. A total of 15 holes (1.5 mm diameter) were drilled radially at different elevations for 

installation of the thermocouples to be used for cooling water temperature measurements. The 

measured cooling water temperature is used to determine heat flux profile along the annulus 

region. The jacket pipe was thermally insulated to reduce environmental heat losses.

6
2.4 Instrumentation

The details of the technical features of the equipment are given in Appendix B. 

2.4.1 Thermocouples 

Thirteen thermocouples were inserted into the holes which have been drilled on the 

outer surface of the stainless steel condenser tube with an angle of 301 and soldered by silver. 

The distance between the inner wall and the tips of thermocouples is approximately 0.5 mm. 

(Figure 2.2) 

Fifteen thermocouples were inserted into the holes, drilled on the outer surface of the 

jacket pipe, and fixed by compression fittings sealed by Teflon material. Thirteen of these 

thermocouples are at the same elevation as the thermocouples to be used for inner wall 

temperature measurements. Besides this, two additional thermocouples were inserted at the 

same elevation but at a 1800 offset orientation. The purpose of these two additional 

temperature measurements is to observe the angular variation of the cooling water 

temperature. 

Ten thermocouples were fixed to a 2 mm diameter Inconel guide wire and installed at 

the central position of the condenser tube for the central temperature measurements. The guide 

wire was fixed at both ends of the test section. 

The installation locations of all thermocouples are given in Figure 2.3.

7
Inner Wall of : 

Condenser Tube ," 30 0 

2.5 mm 

3 mm 

Figure 2.2. Inner Wall Measurement Technique

8

\ i /
Thermocouples 

0101 0

0 

0 

0 

0 

0 

0

100mm 

100mm 

100mm 

200 mm 

200 mm 

200 mm 

200 mm 

200 mm 

200 mm 

200 mm

0

0

0

0

a

0

0

0 

0

0

Condenser Tube

D 

D 

D 

D 

D 

Dl 

ED

Figure 2.3. Installation Locations of the Thermocouples

9

50 mm

100 nnm 

100 mm

-I-

-Ii _

2158 mm

I 

-D 

I

0 

0 

0

Jacket Pipe I r I

-A
All thermocouples for inner wall and cooling water temperature measurements are of 

L-type (Fe-Const type designed according to DIN Standard) and sheathed by Inconel material. 

The thermocouples used for condenser tube central temperature measurements are of J-type 

(Fe-Const type designed according to USA standards) and these thermocouples have a 

temperature-voltage relation very similar to those of L-type thermocouples. The nominal 

outside diameter and wire diameter of all sheated thermocouples is approximately 1.5 mm and 

0.3 mm, respectively. The precision of this type of thermocouples, which belong to second 

tolerance class according to the standard IEC 584-2, is ±2.5 'C between -40 'C and 333 'C. 

The deviation of each thermocouple measurement as compared with the mercury 

thermometer and T-type thermocouple (manufactured by OMEGA) measurements is given in 

Figs. 2.4-2.7 for T,,=50 'C, 97 °C, 150 'C and 180 'C, respectively. As a result of these 

comparisons, maximum deviations from reference measurement are found to be -0.82%, 

-1.16%, -1.13% and +0.83% for nominal temperature settings of 50 'C, 97 'C, 150 'C and 

180 °C, respectively.

50 

49.8 

49.6 

49.4 

49.2 

49 

48.8 

48.6 

48.4 

48.2 

48

1 3 5 7 9 11 13 15 17 19 21 23 25 27 29 31 33 35 37 39 41 

T hermocouple Number 

Figure 2.4. Comparison of Thermocouple and Reference Temperature 

Measurements at Tn=50 °C (Water)

10

i iIjl I 

R TC 

* REF. T
! T 

Pi1 U =.11 

1111:,111:1~~ T 177T TT I -' ý

11 13 15 17 19 21 23 25 27 29 31 33 

T hermocouple Number

35 37 39 41

Figure 2.5. Comparison of Thermocouple and Reference Temperature 

Measurements at TL=97 'C (Water)

I 3 5 7

"~~ JI Il i= " 

I I I , I l 

-1 IL

9 11 13 15 17 19 21 23 25 27 29 31 33 35 37 

T hermocouple Number

Figure 2.6. Comparison of Thermocouple and Reference Temperature 

Measurements at T,=150 °C (Air)

11

98 

97.5 

97 

96.5 

96 

95.5 

95 

94.5 

94

1 3 5 7 9

156 

154 

152 

150 

148 

146 

144 

142

I
I

184 

183 

182 

181 

180 

179 

178 

177 

176 

175

1 3

- TC, 

i~~~ _ LL!it...,~~ _ U lREF.T 

5 7 9 11 13 15 17 19 21 23 25 27 29 31 33 35 37 

T hermocouple Number

Figure 2.7. Comparison of Thermocouple and Reference Temperature 

Measurements at Tn=180 'C (Air) 

2.4.2 Pressure Transducer 

A strain gauge type pressure transducer (product of Transinstruments Inc.), installed at 

the entrance of the test section, can be used for pressure measurement in the interval of 0-6 

bars (g) and has an output of 4-20 mA. The power supply of the transmitter is 24V DC. The 

calibration of this transmitter was made by using air as the operating medium and the result of 

this calibration is shown in Figure 2.8.

12
I

20 

18 

16 

14 

12 

10 

8 

6 

4 

2 

0

0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5.5 

Gauge Pressure (bar) 

Figure 2.8. The Calibration Relation of the Pressure Transducer

6

2.4.3 Flowmeter 

The flow rate measurement of steam is performed by the differential pressure 

transmitter which is the product of ABB Kent-Taylor Company. The device produces current 

in the range of 4-20 mA corresponding to the differential pressure range of the transmitter 

which is 11.7-70 kPa. The current output of the device is linear with respect to differential 

pressure. 

Three types of orifices with flow diameters of 6, 10 and 12.5 mm were calibrated by 

using water as operating medium and flow coefficients, as a function of Reynolds number, 

were obtained. The characteristics of the orifice (flow diameter: 12.5 mm) used in the 

experiments is shown in Figure 2.9. The data obtained during calibration was fitted using the 

following relation:

a = 0.2 * In(Re) - 1.458 (2.1)

where ca is the flow coefficient and Re is Reynolds number.

13
0.6 -. . 

0.55 

0.5 

0.45 W 

0.4 

0.35 X 

0.3 

0.25 

0.2 

0.15 X- Flow Coeff. 

0.1 

0.05 ..... Flow Coeff4lt 

0 

0 5000 10000 15000 20000 25000 30000 

Reynolds Number (--) 

Figure 2.9. The Characteristics of the Orifice with the Diameter of 12.5 mm 

2.5 Data Acquisition System 

The data acquisition system consists of one multifunctional analog and digital I/O card 

(data acquisition card, PCL-812PG), three programmable amplifier and channel multiplexing 

daughter board (PCLD-889) for the analog input channels of the data acquisition card which 

have a total of 48 channels, and an IBM-486 computer system. The data acquisition card and 

multiplexers are the products of Advantech Co., Ltd, and technical details are given in 

Appendix A. 

Total of 40 thermocouple wires was mounted on the multiplexing daughter cards. The 

temperatures were continuously monitored on the colour monitor during experiments. The 

GENIE software was installed to read, display, and log data to disk. The GENIE software, 

which is the product of Advantech Co., Ltd., is designed to run in the Microsoft Windows 95 

environment. GENIE provides an intuitive object oriented graphical use interface that 

simplifies control strategy and display setups. The user can design his own strategy for 

controlling signals during experiments or any kind of industrial process.

14
CHAPTER 3

OPERATING PROCEDURES OF THE TEST FACILITY 

The operating procedure of the METU-CTF consists of two main stages: system check 

and experiments. The former stage is important to understand the response of thermocouples 

at certain operating conditions and to estimate the rate of environmental heat losses. The 

second stage includes system start-up, data logging and system shutdown. 

3.1 System Check 

3.1.1 Isothermal Check of Thermocouples 

The isothermal check of the thermocouples was done at various temperature levels. 

The results of these two tests are given in Table 3.1. The first test (T-20 °C) was performed 

with stagnant water while the second one (T-110 'C) was under high flow rate 

(1.72x 102 kg/s) and high pressure (P-1.4 bars) conditions. In the second test, steam and two

phase flow was established in the condenser tube and jacket pipe, respectively. Apart from the 

isothermal check of the thermocouples, the second test has also shown that the centrally 

located thermocouples were operating properly, at least in the range of thermocouple 

tolerances, when the saturation temperature (T-1 10 °C) was concerned.

15
Table 3.1. Data Collected in the Isothermal Check of Thermocouples 

TC No TC Code Temperature (°C) 

I TC-1 19.4 

2 TC-2 19.8 111.5 

3 TC-3 19.5 110.7 

4 TC-4 19.9 110.3 

5 TC-5 19.9 109.7 

6 TC-6 19.9 111.6 

7 TC-7 19.5 112.4 

8 TC-8 19.9 110.3 

9 TC-9 19.5 110.7 

10 TC-10 19.5 109.8 

11 TC-11 19.9 109.4 

12 TW-1 17.6 106.7 

13 TW-2 17.9 107.9 

14 TW-3 18.0 105.3 

15 TW-4 18.0 109.3 

16 TW-5 18.5 108.5 

17 TW-6 18.5 109.8 

18 TW-7 18.5 107.0 

19 TW-8 18.5 108.9 

20 TW-9 18.0 109.8 

21 TW-10 18.5 109.3 

22 TW-11 18.5 109.7 

23 TW-12 18.0 110.3 

24 TW-13 18.5 109.8 

25 TJ-1 16.6 104.0 

26 TJ-2 16.5 104.4 

27 TJ-3 16.1 105.8 

28 TJ-4 16.1 104.9 

29 TJ-5 17.1 105.3 

30 TJ-6 17.1 104.0 

31 TJ-7 104.3 

32 TJ-8 15.1 103.5 

33 TJ-9 106.1 

34 TJ-10 16.6 103.4 

35 TJ-11 15.6 103.5 

36 TJ-12 16.1 

37 TJ-13 104.3 

38 TJ-4R 15.7 104.8 

39 TJ-12R 104.0 

Notes: TC: central thermocouple 

TW: wall thermocouples 

TJ: jacket thermocouples 

R extension stands for thermocouple at the reverse side 

of the one with the same number

16
3.1.2 Prediction of Environmental Heat Loss

The effect of environmental heat loss on measurements was tested by water under 

stagnant conditions. Both sides of the test section were filled with hot water fed directly from 

the boiler. The data collection duration was 422 s. As can be seen in Figure 3.1, insulation of 

the test section is effective so that rate of temperature decrease was measured to be 

-0.007 °C/s which yields a maximum heat loss of about 0.1 kW at this operating condition

2 

E 

0.) 

0

90 

88 

86 

84 

82 

80 

78 

76 

74 

72 

70

0 50 100 150 200 250 

Time (s)

300 350 400 450

Figure 3.1. Effect of Environmental Heat Loss on Temperature Measurements 

in Jacket Pipe (TJ-4 at 80 cm and TJ-10 at 180 cm, from bottom of test 

section, Tev - 17.6 °C) 

3.1.3 Prediction of Fin Effect for Inner Wall Temperature 

Measurements 

As described in Sub-section 2.4.1, inner wall temperature measurements were 

performed by thirteen thermocouples inserted into the holes which have been drilled on the 

outer surface of the condenser tube with an angle of 30' and soldered by silver. The sheathed 

thermocouple wires are passing through the jacket pipe so that the inner wall temperature 

measurements could be affected by local cooling water temperature by a possible fin effect.

17
A test was performed by filling the condenser tube with stagnant water (Tc-80 'C) and 

by keeping the cooling water (Ti-17 °C) flowing through the jacket pipe at a rate of about 

0.25 kg/s. The flow rate was kept close to the cooling water flow rate values given in the test 

matrix (Chapter 4). The reason for keeping the hot water stagnant inside the condenser tube 

was to check the difference between the bulk and inner wall temperatures. It was expected that 

the difference would be small if fin effect was not dominant. The test result is presented in 

Figure 3.2. It is seen that the temperature difference is about 8 °C and 4 'C at 20 s and 160 s, 

respectively. It is clear that the difference is more at the beginning of the transient process 

during which the system temperature is high. It should also be noted that a temperature profile 

could develop in radial direction due to local natural convection currents so that the measured 

temperature difference between centerline and inner wall seems reasonable. To support this, 

the wall temperature at exactly the same radial distance from the inner surface (0.5 mm) is 

also predicted by the RELAP5/mod3 thermal-hydraulic system analysis computer code [4] by 

imposing the same test boundary conditions, and as illustrated in Figure 3.3, the discrepancy 

was small (+±.2 C). Because of this reason, no correction was made for the inner wall 

measurements.

18
45 70 95 120 145 

Time (s) 

Figure 3.2. Comparison of Centerline and Inner Wall Temperatures 

(Position of Thermocouples: 0.958 m from top)

290.0 ' 

50.0 150.0 250.0

Time (s) 

Figure 3.3. Comparison of Predicted (RELAP5) and Measured 

Inner Wall Temperature

19

90 

80 

70 

60 

50 

40 

30 

20 

10 

0

.2 

o. 

E 

0

20 170

320.0 

S310.0 

0. 

E 

I

0 

C 300.0
3.1.4 Reproducibility of Data

It is important to demonstrate the reproducibility of the data and some experiments 

were repeated at nearly the same operating conditions to see whether the data can be 

reproduced reasonably close. The results of two sets of experiments are presented in Figures 

3.4 and 3.5, for the nominal system pressures of 2 and 3 bars. The experimental boundary 

conditions are given in Chapter 4. The inlet air mass fraction of these experiments is about 

10%. Maximum deviation found from these two sets of experiments are; 6% and 10%, for 

Pn=2 bars and Pn=3 bars, respectively. Primary reasons of deviation are the system pressure, 

vapor flow rate, and cooling water flow rate. However, the percent deviations may be 

considered as reasonable when the uncertainty band of heat flux is considered. The uncertainty 

band of heat flux was calculated to be 11% (Appendix B). 

120000 __ _ S• RR-221 r2 

100000 X_ 

80000 

60000 

40000 

20000 

0I

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 

Axid Posifion (m) 

Figure 3.4. Test of Reproducibility of Experimental Data 

(P,=2 bars, Xi=10 %)

20
160000 

140000 - R-231 120000 -•R-2 .0 

100000 

80000 • 60000 • • 

40000 

20000 

0 , 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 

Axial Position (in) 

Figure 3.5. Test of Reproducibility of Experimental Data 

(Pn=3 bars, Xi=10 %) 

3.2 Experiments 

The experimental procedure comprises the stages of system start-up, transient process 

to steady-state condition, data logging, and system shutdown. 

3.2.1 System Start-up 

First step for the system start-up is the preparation of the boiler by checking water 

inventory. Then system check is made through the control of the following equipment: 

1. Valve position of fresh water inlet of the boiler: closed 

2. Valve position of compressed air inlet to the boiler: closed 

3. Isolation valve of the boiler: closed 

4. Control of boiler pressure controller: set to the predetermined 

pressure level 

5. Control of electrical heaters 

6. Drainage of condensate accumulated inside of the rotameters 

at the exit of the compressor tanks

21
7. Control of pressure inside of compressor tanks 

8. Start of cooling water flow through the jacket pipe 

9. Adjust the flow rate to a predetermined value 

10. Open the valve at the exit of the discharge tank 

11. Open the boiler isolation valve slowly when pressurization is completed 

12. Dump vapor to atmosphere for at least 15 minutes 

13. Check for temperature reading at upstream of the test section whether 

air was swept out by vapor flow. 

14. Adjust system pressure and vapor flow rate by controlling the valve 

at the exit of the discharge tank 

15. Turn on the electrical pre-heaters 

16. Open valve for air injection 

17. Adjust air flow rate via rotameter 

3.2.2 Operating at Steady-state Conditions and Data Logging 

The control of the steady-state conditions was performed by monitoring the system 

parameters, i.e. temperature, pressure, and flow rate, on computer via data acquisition system. 

Data recording was not started before desired steady-state operating condition was reached. 

After steady-state condition was sustained, the data logging was started. Data were recorded 

with one second intervals approximately for a two minutes period. During data logging, 

special care was given to the control of compressed air flow rate, vapor or mixture flow rate, 

and system pressure. 

3.2.3 System Shutdown 

System shutdown was started by putting off the electrical heaters in the boiler and at 

the pre-heating section. Then, compressed air injection was terminated along with fully 

opening of the valve at the exit of the discharge tank. Vapor inside the dome of the boiler was 

discharged till the system pressure was about 1 bar. Finally, the cooling water flow was 

terminated.

22
CHAPTER 4

EXPERIMENTAL TEST MATRIX 

The experimental test matrix consists of two parts: pure steam runs and air/steam 

mixture runs. The following parameters were considered while generating this matrix: 

1- system pressure, 

2- steam flow rate, 

3- air mass fraction. 

The experimental test matrices are given in Table 6.1 and Table 6.2. The coding of 

these experimental runs are based on the following logic: 

RUN-XYZ 

air mass fraction system pressure vapor mass 

1: 0% (pure steam) flow rate 

>1: >0% (air/steam mixture) 

It is to be noted that the mass flow rate measurement was performed by using a 

differential pressure transmitter (orifice meter) so that in each set of experimental run, 

differential pressure, rather than mass flow rate, was set to an almost constant predetermined 

value while changing system pressure. This means, that the vapor mass flow rate increases as 

the system pressure increases (Z= I). In air/steam mixture runs, system pressure and steam 

mass flow rate settings are kept close to those set before during pure steam runs. The reason of 

selecting steam mass flow rate as a fixed parameter, rather than total mixture mass flow rate, 

in air/steam mixture runs is to fix the amount of steam at the entrance of the test section to be 

able to make better comparison with the data of pure steam runs and to understand inhibiting 

effect of air as a noncondensable gas. However, the total mass flow rate (mass flow rate of air 

+ vapor) was increased compared to the pure steam runs and this should be taken into account

23
when analyzing the experimental data. Some additional experimental runs (Z=2) were 

perfomed by using same vapor flow rate and different system pressure: RUN-1.2.2 and 

RUN-1.3.2 based on vapor flow rate of RUN-1.4.1, RUN-3.3.2 based on vapor flow rate of 

RUN-3.4.1R I and RUN-5.3.2 based on vapor flow rate of RUN-5.4.1. Besides these, the total 

mass flow rate (vapor+air) was kept same (Z=3) as the mass flow rate of corresponding pure 

steam run(s) while keeping the system pressure same but varying the air quality only: RUN

3.4.3 and RUN-5.4.3 based on the vapor mass flow rate of RUN-1.4.1, RUN-3.2.3 based on 

the vapor mass flow rate of RUN-1.2.1 and RUN-3.3.3 based on the vapor mass flow rate of 

RUN-1.3.1. 

The effect of inlet superheating was checked by the runs with the extension of 'NH' 

which stands for 'No Heating' at the inlet of the condenser tube. It is to be noted that other 

runs without the extension of 'NH' were performed with certain steam superheating degree at 

the inlet of the condenser tube. The extension 'R' means repeated run with the same inlet 

conditions. However, in some cases, the base runs of repeated ones are not given since those 

runs are not that reliable. 

Table 4.1 Test Matrix for Pure Steam Experimental Runs 

Code P (bars) h, (kg/s) Re, ft (kg/s) Xair 

RUN-I.1.1 1.376 1.409x10-2 43814 0.177 0.0 

RUN-1.2.1 1.829 1.808x10 2 54770 0.221 0.0 

RUN-1.2.1R 1.799 1.812x10 2 54991 0.242 0.0 

RUN-1.3.1 3.029 2.314xl0. 2 66875 0.223 0.0 

RUN-1.4.1 3.959 2.721 x 10-2 76645 0.226 0.0 

RUN-1.5.1 4.837 3.101x10 2 85675 0.225 0.0 

RUN-1.6.1 5.452 3.419x10 2 93365 0.226 0.0 

RUN-1.2.1NH 1.919 1.831x10-2 55236 0.233 0.0 

RUN-1.3.1NH 2.970 2.297x 10.2 66502 0.237 0.0 

RUN-1.4.1NH 3.91 2.701x10.2 76187 0.239 0.0 

RUN-1.2.2 1.919 2.740x10. 2 77183 0.240 0.0 

RUN-1.3.2 3.1 2.800x10"2 80742 0.240 0.0

24
Table 4.2 Test Matrix for Air/Steam Experimental Runs

Code P (bars) mh, (kg/s) Re, ho- (kg/s) Xir 

RUN-2.1.1 1.544 1.465x0o-2 45091 0.199 0.106 

RUN-3.1.1R 1.454 1.530x102 47957 0.229 0.194 

RUN-4.1.1R 1.469 1.7 76x102 55694 0.232 0.211 

RUN-2.2.1 1.919 1.77IX102 53748 0.198 0.099 

RUN-2.2.1R2 1.919 1.853XI02 56228 0.232 0.095 

RUN-3.2.1 1.956 1.865X102 56879 0.236 0.191 

RUN-4.2.1 2.01 2.055X102 62949 0.232 0.275 

RUJN-2.3.1 2.969 2.306XIO2 66771 0.23 0.099 

RUN-2.3.1R1 2.93 2.428X-102 70801 0.232 0.092 

RUN-3.3.1 2.901 2.366X 10-2 69543 0.231 0.189 

RUN-4.3.1 3.16 2.664X 10-2 78258 0.238 0.279 

RUN-5.3.1R 3.13 1.804X10-2 53938 0.237 0.421 

RUN-2.4.1 3.982 2.833X10 2 80253 0.234 0.097 

RUN-3.4.1 3.90 2.77X10-2 79188 0.223 0.193 

RUN-3.4.1R1 3.79 2.644x 10-2 75884 0.237 0.208 

RUN-4.4.1 3.94 2.987X107- 85898 0.231 0.274 

RUN-5.4.1 3.94 2.193X10-2 63663 0.253 0.369 

RUN-6.4.1 3.906 1.526XI1-2 45195 0.253 0.519 

RUN-2.5.1 4.312 2.918XIO12 82043 0.237 0.097 

RUN-6.5.1 4.36 1.881X102 54476 0.253 0.43 

RUN-2.6.1 5.257 3.386X10"2 93388 0.234 0.098 

RUN-2.2.1NH 1.88 1.748x10"2 53269 0.237 0.130 

RUN-2.3.1NH 2.93 2.416x10-2 70480 0.237 0.098 

RUN-2.4.1NH 4.06 2.864x102 81074 0.237 0.113 

RUN-3.4.INH 3.79 2.630x10-2 75697 0.237 0.242 

RUN-4.4.1NH 3.87 2.982x10 2 85980 0.237 0.284 

RUN-5.4.1NH 4.09 2.109x10-2 61278 0.240 0.408 

RUN-3.3.2 2.97 2.749x10"2 80492 0.240 0.168 

RUN-5.3.2 2.97 2.380x10-2 70533 0.241 0.310 

RUN-3.4.3 3.98 2.118x102 60615 0.241 0.230 

RUN-5.4.3 3.94 1.647x102 47968 0.241 0.398 

RUN-3.2.3 1.90 1.276x102 39293 0.250 0.276 

RUN-3.3.3 3.27 1.900x102 55369 0.250 0.223

25
CHAPTER 5

DATA REDUCTION PROCEDURE 

The following steps were used for data reduction: 

A. Averages of all the measured parameters were taken for 10 s and recorded on the computer 

during experiments. The total data logging period of experiments were about 1.5-2 

minutes. The oscillatory behavior was observed only for cooling water temperature with a 

period of 3-4 s, however, the amplitude of the oscillatios was small (±1 C). The source of 

such oscillations observed for the jacket cooling water measurements was turbulence inside 

the annular region of the jacket pipe, as expected. 

B. The following exponential curve was used to fit the cooling water temperature data 

measured inside the jacket pipe, 

T(x) = Ae-B' (5.1) 

Before a decision was given for the type of fitting, different forms were tested, such as 

polynomials, power law, quadratic, and exponential. In general, all these types of fitting 

models yield close results with respect to the correlation coefficient which is defined as 

o r 2 (5.2) 

where Y and S are standard deviation and deviation from the fitting curve. However, 

exponential form is more realistic when the behavior of the heat flux, as the function of the 

axial distance, is concerned. The correlation coefficient was calculated to be between 0.98 

and 1.0, which is quite acceptable.

26
C. Spatial derivative of cooling water temperature T(x) found in step (B) was used to predict 

the heat flux profile in the annulus of jacket pipe. Since the environmental heat loss from 

the jacket pipe is very low no correction was applied to the heat flux profile prediction. The 

local heat flux at the inner tube wall, based on inner diameter (di) of the condenser tube, 

was calculated from

q "(x)= rh'cp dT, (x) 

,rdi dx (5.3)

The derivation of Equation (7.3) was based on the energy balance given in Figure 5.1, for 

steady state and steady flow conditions. 

r$-dx 

Figure 5.1. Energy Balance on a Control Volume of the Jacket Pipe 

The energy balance then yields

dq,, = ih (h.,dX - h.) (5.4)

or

27
7 = r, dh (5.5) 

and by inserting dh=cP dT, we get 

dq, = rhc dT (5.6) 

D. The experimental heat transfer coefficient was then calculated from 

q "(x) 

h(x) = (5.7) 

E. The absolute system pressure was calculated from the measured gauge pressure at the inlet 

of the test section by considering the measured local absolute pressure in Ankara. This 

modification for the measured absolute system pressure was needed to calculate the correct 

saturation temperature of vapor at the inlet of the test section, and partial pressure and 

saturation temperature of vapor along the tube. 

F. The mass flow rate of vapor or air/vapor mixture and associated Reynolds number were 

calculated. For this step, the calibration data of the orifice, i.e. flow coefficient and 

Reynolds number relation given in Section 2.4.3, were used. 

Mass flow rate computation is purely a mathematical process once a calibration curve of 

the orifice has been generated. Since flow coefficient is dependent on Reynolds number, 

which is itself dependent on mass flow rate, final value of the flow coefficient, and hence 

of mass flow rate, can be obtained iteratively using an initial chosen value of Reynolds 

number. 

For the computation of mass flow rate, the method given in BSI-1042 [5] was used. Initial 

guess for the Reynolds number was 106 and the iterative procedure included the calculation

28
of flow coefficient from Equation 2.1, and mass flow rate from the relation of Reynolds 

number based on internal pipe diameter (D) at the upstream of the orifice: 

ReD D (5.8) 

A Dil 

Here rh and g1 are given for vapor or air/vapor mixture, depending on the method of air 

injection, i.e. if air is injected to water inside the boiler then mixture properties should be 

used. 

Then, differential pressure over the orifice as calculated from 

1P Ph (5.9) 

4 

was compared to the one measured and the iteration, by changing the Reynolds number 

guess, continued till the calculated differential pressure was equal to the measured one. The 

density (p) in Equation (5.9) is calculated by considering upstream condition of the orifice 

(vapor or air/vapor mixture depending on the method of air injection).

29
CHAPTER 6

EXPERIMENTAL RESULTS AND DISCUSSION 

6.1 Introduction 

Some of the selected experimental results are presented in this chapter [2,9]. The 

whole range of the data was included into the USNRC Data Bank. The parameters considered 

are; centerline, inner wall and cooling water temperature distributions, heat flux profile, and 

predicted local air mass fraction and film Reynolds number distributions along the flow 

direction. 

6.2 Temperature Distribution 

Temperature measurements were performed at three different locations in the radial 

direction of the test section: at the centerline and inner wall of the condenser tube, and at the 

annulus of the jacket pipe. The centerline temperature simply gives the information for the 

state of vapor, flowing downward, along with the system pressure measured at the inlet of the 

test section. It is expected that for pure steam runs, the measured centerline temperature 

should be the saturation temperature at the corresponding system pressure measured at the 

inlet, by assuming that differential pressure along the channel is small enough (-0.3 bar at the 

system pressure of 5 bars, and much smaller for lower system pressure settings). Moreover, 

when air/vapor mixture flows along the test section, the centerline temperatures indicate the 

existence of air at the core of the condenser tube since the vapor temperature is lower than the 

saturation temperature corresponding to the total system pressure due to partial pressure of 

vapor phase which decreases with increasing quality of air, as Gibbs-Dalton Law states. 

Measured inner wall temperature values also indicate the effect of the presence of air as a 

noncondensable gas, by following the trend of centerline temperature, i.e. higher the 

percentage of air lower the centerline and inner wall temperatures. In fact, air, presumably 

homogeneously mixed with vapor at the entrance of the test section, then tends to accumulate 

at the interface of liquid film and air/vapor mixture which, consequently, causes a 

corresponding reduction of partial pressure of vapor at the interface. In turn, this reduces the

30
saturation temperature at which condensation takes place. The net effect is to lower the 

effective thermal driving force (Ti-Tw) thereby reducing the heat transfer rate. The 

accumulation of air at the interface is the principal reason for the mass diffusion resistance in 

radial direction which causes lower condensation rates. The mechanism of air accumulation at 

the interface of air/vapor and liquid film can be explained on the following physical grounds: 

The vapor that is to be condensed is carried towards the wall of the condenser tube and it also 

carries with it some amount of air. Since the condensate film is impermeable to air, it must be 

removed from the interface at the same rate as it arrives, at steady-state conditions. However, 

the rate of diffusive flow depends on the concentration gradient and sufficient amount of gas 

should be accumulated at the interface to sustain the balance between the convective inflow 

and diffusive back-flow. 

The temperature measurement in the jacket pipe, on the other hand, enables the 

prediction of the local heat flux distribution inside the condenser tube as described in 

Chapter 5. 

The measured temperature distributions of the experimental runs corresponding to the 

nominal system pressure of 2 bars are presented in Figures 6.1-6.3.

180 

160 . 

140 _ 

120 

IO00 

80 _ 

60

20 

0 17 / 

0.00 0.25 0.50 0.75 1.00 1.25 1.50 1.75 2.00 2.25 2.50 

Axial Position (m)

A Tj 

-U-Tw 

-X-Tc 

--- *Tj-fit 

- -Ts

Figure 6.1. Temperature Distribution (P,=2 bars, Re,=54770, Xi=O %)

31

E 

D!
1801 __ _ ___ 

160

140 - t Tj 

120 -- *-Tw 

- X Tc 

S100 -

so

0 I - Tj-fit 

E 80-1 - - - Tc-sat 

6--Ts 

40

20 -

0 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.2. Temperature Distribution (P,,=2 bars, Re,=56228, Xi=10 %) 

160T 

140 

120 -A Tj 

100 -U-Tw 

-80 -X-Tc 

E 60 -_---e---Tj-fit 

w 40 _- - - Tc-sat 

20A _-- Ts 

0

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.3. Temperature Distribution (P,=2 bars, Re,=62949, Xi=28 %)

32
These three experiments were performed at nearly the same operating condition concerning 

the system pressure and inlet superheating degree of vapor. When results given in Figures 6.2 

and 6.3 are compared to those given in Figure 6.1 (pure vapor), following results can be drawn 

at first glance: centerline and condenser tube inner wall temperature distributions are almost 

constant throughout the tube length regardless of the inlet air quality. The centerline, wall and 

cooling water temperatures decrease as inlet air quality increases. As can be seen in Figure 

6.1, the centerline and wall temperatures are very close to the saturation temperature (around 

120 °C), corresponding to the system pressure, especially at the entrance of test section. It is to 

be noted that the rate of condensation, along with the cooling water and vapor mass flow rates, 

becomes important regarding the rate of detachment of the liquid droplets from the condensed 

liquid film over the tube surface towards the thermocouples which are attached to a guide wire 

at the centerline of the tube. That is to say, when the rate of condensation increases the 

measured centerline temperatures tend to decrease (centerline temperature subcooling effect) 

due to higher rate of detachment of liquid droplets or patches from the surface. In the test 

matrix under consideration, the aforementioned effect is more pronounced as system pressure 

and vapor flow rate are increased, i.e. the former is effective for higher heat flux whereas the 

later is responsible for agitating turbulence of liquid film on the inner surface of the condenser 

tube. If, somehow, the operating condition reverses then the inlet superheating of the vapor 

becomes more sensible, as seen in Figure 6.1. 

In Figures 6.2 and 6.3, it is observed that, difference between the saturation 

temperature (To-sat), corresponding to the system pressure measured at the inlet of the test 

section, and the measured centerline temperature (TJ) increases as the inlet air quality 

increases, i.e. on the average, AT=11-17 'C and 19-27 'C, for Xi=10 % and 28 %, 

respectively. In reality, the saturation temperature also decreases as the local mass fraction of 

air increases and this is shown in Figures 6.2 and 6.3 by the parameter 'T,'. T, was calculated 

from the predicted local vapor pressure given by the Gibbs-Dalton ideal gas mixture equation: 

, P, X - 1) (6.1) 

1- 2Xar j - I1

33
where the local air mass fraction (Xair) was predicted from the sectionwise energy balance. 

Difference between the saturation temperature of pure vapor (T-sat) and predicted 

saturation temperature of air/vapor mixture (Ts) is about 3 'C for the case with X=-10 %, and 

7 'C for the case with Xi=28 %. The marked increase of the temperature difference (Tc-sat-Ts) 

in two cases as presented in Figures 6.2 and 6.3 is the result of increase of the inlet air mass 

fraction, as expected. Another point worth to mention is that an inequality between T, and T, 

(T, < T) is observed for all cases, except the pure vapor case, and the difference of T, and T, 

increases from 9-14 'C to 13-19 °C by an increase of air mass fraction from 10 % to 28 %. 

This can be attributed to the lowered inner wall temperature due to higher air mass fraction, 

which in turn lowers the measured centerline temperature as well. This situation indicates that 

the measured centerline temperature was highly affected by the prevailing inner wall thermal 

conditions, probably, due to detached liquid droplets which lowers the measured centerline 

temperature (T,). This is important since this situation, i.e. having inequality between 

measured and predicted centerline temperatures (Tr < Ts), leads to the conclusion that pattern 

of the core flow inside the condenser tube is likely to be homogeneous two-phase flow mixed 

with air. It should be noted that two-phase flow condition inside the tube makes it difficult to 

predict the condensation heat transfer coefficient since selecting either of centerline 

temperatures, that is the measured or the predicted one, for calculating the heat transfer 

coefficient may lead to considerably different local values. 

The measured temperature distributions of the experimental runs corresponding to the 

nominal system pressure of 4 bars are presented in Figures 6.4-6.7. All of the temperatures are 

elevated compared to the previously discussed case due to higher system pressure and inlet 

temperature. It should be noted that experimental runs with the system pressure of 4 bars were 

performed at the condition of higher vapor Reynolds number compared to the case with 2 bars, 

i.e. the vapor Reynolds number is about 50 % higher than the case with 2 bars system pressure.

34
180 

160 X X IT 

140 -•-X;ý- •=-T--=-'X--X-X--• •• 

1420 - A Tj .-Ui-Tw 

100

s-x-TTc 

60 __oT.-Tj-flt 

40-- -Ts 

20, 

0 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.4. Temperature Distribution (P,=4 bars, Rev=76645, Xi=O %) 

1.80 

160 , 1 

140 ------- ----- ------ I-- A Tj 

1L20 __ __ I_ 

100 -_-x-Tc 

so 1e---Tj-fit ---

60 -.. Tc-sat 

40- -- -Ts 

200 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.5. Temperature Distribution (P,=4 bars, Re,=80253, Xi=10 %)

35
180 

160 

140 

120 

100 

80 

60 

40 

20 

0

4 ____ _____ ___-- _ mi l_. i __--. _______ __i ___ _______ 

-_:- • - =_ 

T--( -x X-X -XX 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m)

A Tj 

--- Tw 

-x-Tc 

--- *--Tj-fit 

-----. Tc-sat

Figure 6.7. Temperature Distribution (P,,4 bars, Re,-45195, Xi=52 %)

36

. .... - A Tj 

. . .-Tw 

-X-Tc 

---- T j-fit 

- - - Tc-sat 

-- Ts 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.6. Temperature Distribution (P,,= bars, Re,=85898, Xi=28 %)

10

180 

160 

140 

120 

100 

80 

60 

40 

20 

0
The pure vapor run results show that the measured centerline temperature (TJ) trend 

closely follows the saturation temperature (TS) line and always remains below of it, with a 

maximum difference of about 4 'C, even at the bottom of the test section. It is observed from 

the comparison of pure vapor runs with the nominal system pressure of 2 bars and 4 bars, that 

the absolute temperature difference (T,-T.) decreases by increase of Re, and the effect of inlet 

superheating is no longer observable for the case with the pressure of 4 bars. This may be the 

result of subcooling at the centerline of the condenser tube due to increased rate of detachment 

of liquid droplets from condensate film. The temperature differences of Tc.sat (saturation 

temperature of zero air quality) and Tc are: 9-12 'C, 24-29 'C and 42-48 °C for the cases of 

Xj= 10 %, 28 % and 52 %, respectively. The differences between the predicted (from the 

Gibbs-Dalton Law, T,) and the measured (TJ) centerline temperatures come out to be 1-5 'C, 

5-9 'C, 13-21 'C, and 15-31 °C, for Xi=0 %, 10 %, 28 %, and 52 %, respectively. The 

temperature difference (Ts.t-Tj) can be attributed to the existence of air and cooling of 

centerline thermocouples by liquid droplets. Among the results presented, the experimental 

results pertaining to the case with Xi=52 % represents a special case concerning the inlet vapor 

Reynolds number (-41000), which is about half of other runs at the same pressure setting. The 

distribution of Ts shows a marked decrease towards the bottom of tube due to increase in X, 

steeper than other runs. Moreover, a sharp decrease in all measured temperatures (Ta, T,) are 

observed for the case with Xi=52 %, i.e. about 40 °C, relative to the pure vapor case. 

6.3 Local Heat Flux Distribution 

To calculate local heat transfer coefficients, the local air/vapor mixture temperature, 

local inner wall temperature, and the local heat flux must be known. The local air/vapor 

mixture and inner wall temperatures were measured directly, and the local heat flux was 

obtained from the measured coolant temperature profile. Hence, the local axial temperature 

gradient (dTcw/dx) was computed from an exponential fit of the measured coolant temperature 

as a function of axial distance and the local heat flux was determined from: 

q"(x)=-hmcP dT(x) (6.2) 

Ird i dx

37
The heat flux distributions for experimental runs corresponding to the nominal system 

pressures of 2-6 bars, and including pure vapor and different mixtures of air and vapor, are 

presented in Figs. 6.8-6.12 (X in these figures stands for inlet air mass fraction).

160000 

140000 

120000 

100000 

80000 

60000 

40000 

20000 

0

0 0.25 0.5 0.75 1 1.25 1.5 1.75 

Axial Position (m)

2 2.25 2.5

Figure 6.8. Heat Flux Distribution along the Condenser Tube 

(Pn=2 bar, Re,=54OOO-63000)

38
250000

200000 x "IX - . .. . ... 

--- X =19% 

150000 ___ 

1000 0 

50000.... 

0 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.9. Heat Flux Distribution along the Condenser Tube 

(P.=3 bars, Re,=67000-78000 and Re,= 54000 for Xi=42%)

250000 

225000 

200000 

175000 

150000 

125000 

100000 

75000 

50000 

25000 

0

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 

Axial Position (m)

2.25 2.5

Figure 6.10. Heat Flux Distribution along the Condenser Tube 

(Pn=4 bar, Re,=77000-86000 and Re,-45000 for Xi=52 %)

39
300000 

1 

25000 -xpure 

100000 -=1"% 

50000 

0 0.25 0.5 0.75 1 1.25 1-5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.11. Heat Flux Distribution along the Condenser Tube 

(P,=5 bar, Re,=82000-86000 and Re,=55000 for Xi=43%) 

350000 1 1 

--x- pure 

300000 Xx -< 

-.- X =d.% 

250000 - ___,,_ 

200000 ___,___ ___ 

150000 -___ 

100000 x 

50000- _ _ _ 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.12. Heat Flux Distribution along the Condenser Tube 

(P,--6 bars, Re,=93000)

40
It is clear in these figures that the heat flux drastically decreases as inlet air mass 

fraction increases. This situation is the evidence for how some amount of air, mixed with 

vapor, degrades the performance of the heat exchanger. By referring to Figures 6.8-6.12, 

decrease in local heat flux at the middle of the test section (-1 m from the top) as compared to 

the corresponding pure vapor case is summarized as: 

- Pn=2 bar: 20 % (Xi=10 %), 24 % (Xi=20 %), 45 % (Xi=30 %) 

- Pn=3 bar: 19 % (Xi=10 %), 24 % (Xi=19 %), 30 % (Xi=28 %), 48 % (Xi=42 %) 

- Pn=4 bar: 22 % (Xi=10 %), 24 % (Xi=20 %), 28 % (Xi=29 %), 37 % (Xi=37 %) 

44 % (Xi=52 %) 

- Pn=5 bar: 24 % (Xi=10 %), 75 % (Xi=43 %) 

- Pn=6 bar: 27 % (Xi=10 %) 

Another point to be emphasized is that local heat flux values for pure steam and 

air/steam mixture runs get closer towards the bottom of the condenser tube due to diminishing 

condensation rate as the result of increased resistance of condensate film. This means that 

condensate film resistance in pure steam runs tends to dominate over diffusion resistance in 

air/steam mixture runs, at the bottom of the condenser tube. 

An increase in system pressure increases local heat flux and this can be attributed to 

the increase in wall subcooling degree that enhances the thermal driving force for heat transfer. 

Moreover, higher system pressure associated with the higher inlet temperature leads to a 

greater number of molecular collisions helping in the diffusive transport of energy. However, 

in our experimental investigation, the dependency of the wall subcooling degree, either 

measured (Tc-Tw) or predicted from Gibbs-Dalton Law (Ts-T,), on system pressure is such 

that the wall subcooling degree remains nearly the same for the same inlet air mass fraction 

and for the different system pressure. This implies that the vapor mass flow rate may dominate 

over system pressure, concerning the effect on local heat flux, for cases with air/vapor mixture 

(Figure 6.14). The situation is rather different in pure vapor runs, that is increase in system 

pressure has a strong effect on enhancement of predicted, and even measured, wall subcooling 

degree and hence on increase of local heat flux (Figure 6.13).

41
250000

200000 

E 150000 

100000 

z 

50000 

0

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m) 

Figure 6.13. Effect of System Pressure (Pure Steam)

160000 

140000 

L20000 

100000 

60000 

40000 

20000 

0

x• - P-x-P:=4 bar, 

"Re=75697 

.X -- - ---P=3 bar, 

"• -"Re=80492 

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 2.5 

Axial Position (m)

Figure 6.14. Effect of System Pressure (Air/Steam Mixture; Xi=20 %)

42
Total of 9 experiments were repeated by zero inlet superheating degree, as described in 

Chapter 4. From these nine cases, three of them belong to the pure steam cases and the rest of 

them are for air/steam mixture with different air inlet quality settings. The overall finding of 

this investigation reveals the fact that the inlet superheating of steam has no considerable 

effect on heat flux. A sample result for heat flux, corresponding to the case of Pn=4 bars and 

Xi=10 %, is presented in Figures 6.15.

180000 

160000 

140000 

120000 

100000 

80000 

60000 

40000 

20000 

0

0 0.25 0.5 0.75 1 1.25 1.5 1.75 

Axial Position (W)

2 2.25 2.5

Figure 6.15. Effect of Inlet Superheating of Steam (Pn=4 bars, Xi=10 %) 

It should be pointed out that it is not surprising to have no superheating effect on local 

heat flux values since there is a subcooling at the centerline of condenser tube due to possible 

detachment of liquid droplets from the inner surface of tube. Hence, there is no considerable 

effect of dryness of steam at the entrance since it gets wetted throughout its flow in downward 

direction.

43
6.4 Local Air Mass Fraction Distribution

The local air mass fraction was predicted by using the data of predicted heat flux 

distribution. The local air mass fraction is defined as: 

Xair (X) = .ir (6.3) 

?ha + rhm(x) 

To obtain the local air mass fraction (or air quality), the data for the local vapor mass 

flow rate changing along the channel due to condensation are needed. To calculate the local 

vapor mass flow rate, first the sectional condensate flow rate was obtained from a sectionwise 

steady state heat balance by neglecting the gas phase sensible heat transfer and considering 

only the latent heat transfer, which is given as 

Aq = Ahizcod hfg (6.4) 

Here the latent heat of condensation (hfg) was calculated at the measured local inner 

wall temperature since film temperature is much closer to the inner wall temperature than that 

of bulk temperature. The local condensate flow rate, then, was calculated by summing up the 

incremental values up to that point. The local vapor mass flow rate was calculated by 

subtracting the local condensate mass flow rate from the known vapor flow rate measured at 

the inlet of the test section. Since the air mass flow rate is constant, Equation (6.3) can be used 

for determining the local air mass fraction. 

The predicted data for the air mass fraction distribution along the channel helps us 

understand the trend of vapor mass flow rate that decreases towards the bottom of the channel, 

along with phase change due to condensation process. The predicted results of the air mass 

fraction are presented in Figures 6.16 and 6.17 which correspond to the system pressure of 

2 bars and 4 bars, respectively.

44
0.4 

0.3 

S0.2 

0.1 

0

0 0.2 0.5 0.7 1 1.2 1.5 1.7 2 2.2 2.5 

5 5 5 5 5 

Axial Position (m) 

Figure 6.16. Distribution of Air Mass Fraction (P,=2 bars)

0.8 

0.7 

0.6 

0.5 

0.4 

0.3 

0.2 

0.1 

0

2r 

. .. .- XX • t X X-X .--

.Xt-X-l.X-- Xi•4

0 0.25 0.5 0.75 1 1.25 1.5 1.75 2 2.25 

Axial Position (m)

-x- 10% 

- 1W -20% 

--- o-30% 

- -)K- -37% 

•--e.-- 52%

2.5

Figure 6.17. Distribution of Air Mass Fraction (Pn=4 bars)

45
It is seen from the predicted data that the air mass fraction increases steadily along the 

channel as a result of condensate accumulation. This is known as "suction effect" since vapor

liquid phase change due to condensation sucks air towards the bottom of the test section. We 

can extract a general conclusion that the air quality is a linear function of the axial distance in 

spite of the heat flux behavior which exhibits an exponential trend. 

6.5 Condensate Film Reynolds Number 

When there is a sufficient amount of condensate flow, turbulence may appear in the 

condensate film which results in higher heat transfer rates. The criterion for determining 

whether the flow is laminar or turbulent is the Reynolds number, and for the condensation 

system it is defined as 

VDhpf 4AVpf Re =- =- (6.5) 

and by inserting P---Kd for vertical tube, we get 

Ref - (6.6) irdlgi 

where rhn is the mass flow rate through the particular section of the condensate film. 

As the condensate layer thickness increases it can undergo a transition from laminar to 

turbulent flow. McAdams [10] suggests that transition could occur at a condensate Reynolds 

number of 1800 where the Reynolds number is defined in Eq. (6.5). At high values of the 

shear stress, however, Carpenter and Colburn found transition values as low as 200-300 [11]. 

Since Reynolds number of vapor is high (-54000-94000) in the experiments conducted at the 

METU-CTF, the later values of the condensate Reynolds number (200-300) are applicable for 

criteria of transition from laminar to turbulent flow. The sectional condensate flow rate was 

obtained from a sectionwise steady state heat balance and, then, the local condensate flow rate

46
was calculated by summing up the incremental values up to that point. The local values 

of the condensate film Reynolds number were calculated and the results are presented in 

Figures 6.18 for system pressures of 2 bars. The calculated film Reynolds number decreases as 

air mass fraction increases at the same system pressure setting, and falls into the range of 

turbulent region for all experimental runs (other than those performed at a system pressure of 

about 1.5 bars), i.e. Ref > 300. For example; the film Reynolds number increases up to about 

3500 for pure vapor run performed at the nominal system pressure of 6 bars.

1800 

1600 

1400 

1200 

1000 

80 

600 

400 

200 

0

0 0.2 0.4 0.6 0.8 

x/L

1 1.2

Figure 6.18. Distribution of Condensate Film Reynolds Number (Pn=2 bars) 

6.6. Comparison with Theory 

A set of simulations was performed by using the RELAP5/mod-3.2.1.2 thermal

hydraulic system analysis computer code. The code solves six equations (mass, momentum, 

and energy) for two-phase and one equation (mass) for noncondensable gas. The heat transfer 

solution scheme of the code also includes condensation of water vapor containing 

noncondensable gas, such as air, hydrogen, nitrogen, helium. The RELAP5 code calculates a 

wall heat transfer coefficient based on condensation logic under the following conditions:

47
- The wall temperature is below the saturation temperature based on the bulk partial pressure 

of vapor calculated by using the Gibbs-Dalton ideal gas mixture equation. 

- The liquid temperature is above the wall temperature. 

- The liquid void fraction is greater than 0.1. 

- The bulk noncondensable quality is less than 0.999. 

- The pressure is below the critical pressure. 

- Only filmwise condensation exists. 

The default model used currently is the Shah-Colburn-Hougen model. The Shah model 

replaces the Nusselt model for pure steam condensation if the heat transfer coefficient 

calculated by this model is greater than that of Nusselt model. The Colburn-Hougen diffusion 

model, used for taking into account the inhibiting effect of a noncondensable gas, involves an 

iteration process to solve for the steam saturation temperature at the interface between the 

steam/gas boundary layer and water film. The model was developed under the following 

assumptions: 

- The sensible heat transfer through the diffusion layer to the interface is negligible. 

- Stratification of the gas in vapor by buoyancy effects is negligible. 

- Required mass transfer coefficients can be obtained by applying the analogy between the 

heat and mass transfer. 

- The gas is not removed from the vapor region by dissolving it in the condensate. 

The formulation is based on the principle that the heat transferred by condensing vapor 

at liquid-vapor interface which is diffusing through the noncondensable gas is equal to the 

heat transfer through the condensate. This energy conservation principle, obviously, needs 

interface pressure (or temperature) to be determined since the interface pressure is always 

lower than the bulk pressure (total pressure) at the core due to existence of the 

noncondensable gas at the interface. One of the deficiency of the model is the assumption of 

having same gas and vapor velocities which is not necessarily correct in reality. Moreover, the 

effect of superheating of steam was not incorporated in the formulation.

48
Some of the experimental cases with air/steam mixture were simulated by the 

RELAP5/mod-3.2.1.2 beta test version by imposing the measured inner wall temperatures as 

boundary condition which means that the jacket pipe was not modelled for not to increase the 

uncertainty associated with the flow inside the jacket pipe. The results of heat flux predictions 

are plotted against the experimental data as given in Figure 6.19.

200000 

150000 

100000 

50000 

0

0 50000 100000 150000 

Heat Flux-RELAP5 (W/m2)

200000

Figure 6.19. Comparison of Measured and Predicted Local Heat Flux Values 

It is clear that the RELAP5 code overpredicted majority of experimental local heat flux 

data by 5 % - 50 %. In general, the deviation increases by increase of inlet air mass fraction.

49
LIST OF REFERENCES

1. Tanrikut, A., Heper, H., Bayraktar, N. and Gunel, I., "The Simulation of Loss of Heat 

Residual System after Reactor Shutdown," Annual Meeting on Nuclear Technology '94, 

Stuttgart, 1994. 

2. Tanrikut, A., "In-Tube Condensation in the Presence of Air," Ph.D. Thesis, Mechanical 

Engineering Department, Middle East Technical University, Ankara, 1998. 

3. Tanrikut, A., "An Assessment of RELAP5 Code for Pure Steam Condensation and 

Condensation in the Presence of Air," Annual Meeting for Nuclear Technology, Mannheim 

(Germany), 21-24 May 1996. 

4. RELAP5/MOD3 Code Manual, Code Structure, System Models and Solution Methods, 

Idaho National Engineering Laboratory, NUREG/CR-5535, Vol. 1, 1995. 

5. Methods of Measurements of Fluid in Closed Conduits, British Standard Institution, BS

1042, Section 1.1, 1981. 

6. Brinkworth, B. J., An Introduction to Experimentation, The English Universities Press Ltd., 

1968. 

7. Kuhn, S. Z., Schrock, V. E. and Peterson, P. F., "Final Report on U. C. Berkeley Single 

Tube Condensation Studies," Dept. of Nuclear Eng., UCB-NE-4201, 1994. 

8. Siddique, M., Golay, M. W. and Kazimi, M. S., "The Effect of Noncondensable Gases on 

Steam Condensation under Forced Convection Conditions," Dept. of Nuclear Eng., MIT, 

MIT-ANP-TR-010, 1992. 

9. Tanrikut, A., Yesin, 0., "An Experimental Research on In-tube Condensation in the 

Presence of Air," 2nd International Symposium on Two-phase Flow and Experimentation, 

Pisa (Italy), 23-26 May 1999. 

10. W. H. McAdams, Heat Transmission, McGraw-Hill Book Company, Inc., New York, 

1954. 

11. W. H. Rohsenow, J. H. Webber and A. T. Ling, Effect of Vapor Velocity on Laminar and 

Turbulent Film Condensation, Trans. ASME, Vol. 78, 1956.

50
APPENDIX A

PHOTOGRAPHS OF THE METU-CTF

Figure A. 1. General View of the METU-CTF

51
/-

It~ *

Figure A.2. The Compressed Air Supply System 

F U -

Figure A.3. The Upper Part of the Test Section (without the Jacket Pipe)

52
Figure A.4. The Thermocouple (D=1.5 mm) Fixed to the Guide Wire (D=2 mm) to be 

Placed Inside the Condenser Tube for Central Temperature Measurement

Figure A.5. The Thermocouple (D=1.5 mm) Fixed Inside the Condenser Tube for Inner Wall 

Temperature Measurement

53
Figure A.6. The Lower Flange of Test Section and Penetration of Thermocouples Used for 

Central Temperature Measurement

54
APPENDIX B

SPECIFICATIONS OF INSTRUMENTATION AND 

DATA ACQUISITION SYSTEMS 

1) Thermocouples: 

Manufacturer: ELIMKO Co., Turkey 

Type: L (Fe-Const type designed according to German DIN standard) 

J (Fe-Const type designed according to USA standard) 

Class: 2 

Temperature Range: -40 °C to 333 'C 

Tolerance Value: ±2.5 'C according to the standard IEC 584-2 

2) Pressure Transducer: 

Manufacturer: Transinstruments Inc., England 

Type: Strain gauge 

Working Medium: Water, steam, gas 

Pressure Range: 0-6 bar (g) 

Output Signal: 4-20 mA (linear) 

3) Flowmeter: 

Manufacturer: ABB Kent-Taylor, Italy 

Type: Differential pressure transmitter 

Working Medium: Water, steam, gas 

dP Range: 11.7-70 kPa 

Output Signal: 4-20 mA (linear) 

4) Data Acquisition System: 

Manufacturer: Advantech Co., Ltd. 

Data Acquisition Card: PCL-812PG 

Analog Input: 16 single-ended channels 

12 bits resolution 

±5V, ±2.5V, ±1.25V, ±0.625V, and ±0.3125V input range 

0.0 15% of reading ±1 bit accuracy 

Analog Output: 2 channels 

12 bits resolution 

Amplifier/Multiplexer Board: PCLD-889 

Input Channel: 16 

Input Range: ±1OV maximum 

Cold Junction Compensation: +24.4V/°C 

0.OV at 0.0 °C

55
APPENDIX C

ERROR ANALYSIS 

The total error of a function F with independent measured variables xi, X2,, X3 ... xn, 

was obtained as [6]:

F < + • 2+ ............. 

The relative error can be found by dividing the expression given above by F:

UF ... 2 + ( 2+ ............... ( x.2 +11/2 

The experimental heat flux is defined as: 

) ,cc dT,.,(x)

-rai ax

Therefore:

q_"= cp dT,(x) 

Acý Trdi dx

56

(C.1)

(C.2) 

(C.3)

(C.4)
-dq" h, cnp (C.5) 

d(dTc. / dx) MC. 

Substituting Equations (C.4) and (C.5) into Equation (C.1) and dividing both sides of by (q,) 2 

we find the relative error for q": 

__ =( _ 2dr) 2(C.6) 

The variables cp, and d are assumed to be error free. 

Similarly, the relative error for the condensation heat transfer coefficient is found from the 

following equation: 

hexp (W) q"(x) (C.7) 

(TI(x) - TW(x)) 

The relative error for hlp is: 

____ F(O~ (7(T _____) 

= ),-/ )2 (C.8) 

Since we did not use a flow meter for cooling water flow measurement, the relative 

error for measured flow rate is taken from the maximum weight deviation calculated from the 

measurement performed in the beginning and end of each run. It is found that the maximum 

relative error associated with the cooling water measurement is 0.05, on the average, which 

corresponds to a deviation of 200 gr at the 4300 gr total weight of cooling water collected for 

a period of 15 s.

57
The slope of the coolant water axial temperature profile, i.e. dTcw/dx, was determined 

from an exponential fit of the measured coolant temperatures as the function of the 

measurement distance, and for almost all the runs the R2 value for the fit of the data to the 

exponential relation was greater than 0.98. This means that error associated to the curve fitting 

is very small. However, it is difficult to predict the error associated with the temperature 

gradient directly. Only an estimation was made by considering the effect of measured 

temperature data with higher deviation than the general trend of the temperature distribution. 

It was found that the effect of measured temperature data with high deviation yields an error 

of ± 10 %. The percent error of the temperature gradient assumed for the UCB-4 data was 

also ± 10 % [7]. 

For the prediction of the uncertainty associated with the heat transfer coefficient, the 

standard deviation of the thermocouples must be known. The experimental data of isothermal 

check of thermocouples were used for determining the standard deviation of thermocouples. 

The experimental data, as given in Table 3.1, yields a maximum standard deviation of 0.762. 

Since this is an experimental result, the calculated standard deviation includes the tolerances 

given by the factory. 

The aforementioned values of standard deviations and data needed for Equations (C.6) 

and (C.8) are summarized as follows: 

Ym --0.05 rn, (C.9) 

(YTS = YTw = 0.762 (C. 10) 

C(dr/dx) = 0.10 (dT,,/dx) (C.11) 

(Ts-Tw) = 5 °C (C. 12) 

o(Tg-Tw) = (O'Ts+O'Tw) 1 2 (C. 13) 

Substituting values from Equations (C.9) to (C. 13) in Equations (C.6) and (C.8) we get:

58
[6q-] 0.11 

and 

[hma 0.24 

Therefore the maximum uncertainties associated with the heat flux and the heat transfer 

coefficient are ±11 % and ±24 %, respectively. The temperature difference (T,-Tw) was equal 

to or greater than 5 °C in the major part of the condenser tube length in all experiments. 

However, this temperature difference is less than 5 'C in entrance region (-0.25 m) and the 

uncertainty associated with the heat transfer coefficient escalates to ±38 % when a 

temperature difference of 3 'C is assumed. It is also observed that the error associated with the 

heat flux and the heat transfer coefficient increases when the coolant temperature change per 

unit length decreases which happens towards the end of the condenser tube. If we assume a 

conservative value for the deviation of the coolant temperature gradient such as 15 % (instead 

of 10 %), the uncertainties become 16 % and 27 % for heat flux and heat transfer coefficient, 

respectively. This reveals that the uncertainty band increases at the entrance region much more 

than the end of the test section.

59
NRC FORM 336 U.S. NUCLEAR REGULATORY COMMISSION 1. REPORT NUMBER 

(2.-9) (Asned by NRC. Add VoL, Supp., Rev.. 

NRCM 1102. nd Addendum Numbe., It any.) 

3201.3202 BIB OGRAPHC DATA SHEET 

(See minsuconh an h reverse) 

2. TITLE AND SUBTITLE NUREG/IA-0184 

In-Tube Steam Condensation in the Presence of Air 

3. DATE REPORT PUBLISHED 

June 2000 

4. FIN OR GRANT NUMBER 

5. AUTHOR(S) 6. TYPE OF REPORT 

A Tanrikut, 0. Yesin Technical 

7. PERIOD COVERED Mn*si Dwie) 

8. PERFORMING ORGANIZATION - NAME AND ADDRESS (YNRC, provide Division, Ofts or Region, U.S. Nudew Regutatay Cww=s=, and mailing addess, ffoonra 

-rovide erwe nd rai/lg addmss.) 

Turkish Atomic Energy Authority 

Eskisehir Yolu 

06530 Ankara, Turkey 

9. SPONSORING ORGANIZATION - NAME AND ADORESS (tf NRC, type 'Same &s aboved dconfclor, provide NRC Division, Offie or Region, U.S. Nucow Regultoy Cammsib 

aid medliug address.)

Division of Systems Analysis and Regulatory Effectiveness 

Office of Nuclear Regulatory Research 

U.S. Nuclear Regulatory Commission 

Washington, DC 20555-0001

10. SUPPLEMENTARY NOTES

11. ABSTRACT r" words ar les) 

In this research work, in-tube condensation in the presence of air is investigated experimentally for different operating conditions, 

and inhibiting effect of air is analyzed by comparing the experimental data of air/steam mixture with the data of corresponding pure 

steam cases, with respect to temperature, heat flux, air masses fraction, and film Reynolds number. The test matrix covers the 

range of Pn=2-6 bars, Rev=45,000-94,000, and Xj+0% - 52%. The inhibiting effect of air manifests itself as a remarkable 

decrease in centerline temperature (10 oC - 50 oC), depending on inlet air mass fraction. However, the measured centerline 

temperature is suppressed compared to the predicted one, from the Gibbs-Dalton law, which indicated that the centerline 

temperature measurements are highly affected by inner wall thermal conditions, possibly due to narrow channel and high vapor 

Reynolds number.

12. KEY WORDSIDESCRIPTORS 1ist words orpiases &at wo assist remeennw in locating fth repor) 13. AVAILABAIlY STATEMENT 

Condensation unlimited 

Noncondensables 14. SECURrIY CLASSICA'ION 

(rid Page) 

unclassified 

Ob9s Report) 

unclassified 

15. NUMBER OF PAGES 

16. PRICE

NRC FORM 335 (2-89)
Federal Recycling Program
NUREG/IA-0184 IN-TUBE STEAM CONDENSATION IN THE PRESENCE OF AIR

UNITED STATES 

NUCLEAR REGULATORY COMMISSION 

WASHINGTON, D.C. 20555-0001

years

JUNE 2000