Challenger Center Cooling Tower
University of Tennessee at Chattanooga
ENCH/ENEV 435
Lab Report By: Dianah Dugan
Lab Partner: Alex Saputa
Professor: Dr. Jim Henry
Date: October, 8, 2008
ABSTRACT
Over the course of two weeks, two experiments were performed to determine the
heat absorbed by air flow for the Challenger Center cooling tower. This document
illustrates the results from measurements taken across the intake and exhaust air flow.
Temperatures ranged from 98-111°F at the exhaust fans, and 90-96°F at the intake.
Relative humidity ranged from 39-46% at the intake; the total heat absorbed for unit A
was calculated to be 390,000 ± 93,000 Btu/hour, or 33 ± 7 tons of cooling.
2
TABLE OF CONTENTS
Introduction
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Theory .
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Equipment
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Procedure
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Results
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Discussion of Results .
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Conclusions
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Recommendations
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References
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Appendix
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INTRODUCTION
Located behind the UTC Challenger Center, this dry cooling tower works to cool
the entire 23,000 square foot building by two systems, A and B, which do not function
simultaneously. Each system contains four exhaust fans and eight intakes. A side-view
schematic of the cooling tower is shown below in Figure 1.
Exhaust B
Exhaust A
Inside
Intake
Outside
Intake B
Inner
Intake B
Outside
Intake A
Inner
Intake A
Figure 1: Side View Schematic of Cooling Tower
Cool water enters the building, absorbing heat from inside. When the water
becomes warm, it is sent back to the cooling tower where refrigerant absorbs the heat,
cooling the water once again. Refrigerant is then cooled by air flow which is measured
during this experiment. The cooling and warming effects between the air, water, and
refrigerant demonstrate principles in heat exchange.
The purpose of this experiment was to determine a total energy balance for the
Challenger Center cooling tower. In order to do so, temperature, velocity, and relative
humidity readings were collected using specific devices which enabled the mass flow
rate, volumetric flow rate, enthalpy, and absolute humidity to be calculated. The results
of this experiment will then be compared to the solar radiation that the building receives
on an average daily basis.
4
Two experiments were performed over the course of two weeks. It should be
noted that for the first experiment, only outlet temperature and velocity were recorded for
the system. During the second experiment, temperature, velocity, and relative humidity
were recorded at the inlet and outlet for the system. Analysis of the system in terms of
energy, enthalpy, and absolute humidity all come from the second experiment. The first
experiment is used to compare changes in velocity and temperature within the system.
5
THEORY
For purposes of analysis, a pressure of 1 atmosphere, and no work were assumed.
The change is enthalpy of air will be greater than zero, while the change in enthalpy for
water will be less than zero. This is due to the heat exchange process as explained
previously. A complete block diagram of the cooling tower is shown in Figure 2.
Figure 2: Block diagram of Cooling Tower
Through measurements, temperature at the inlet and outlet were found. Velocities
were measured at specific points across four different fans at the air outlet. Wet bulb
temperatures were obtained at the air inlet. Based on these measurements, area can first
be calculated. Once area has been determined, the volumetric flow rate can be solved for,
using equation (2) shown below, where “A” is the area and “v” is the average velocity
measured.
V  A * vavg
(2)
The mass flow rate can now be solved for, as demonstrated in equation (3), by
multiplying the volumetric flow rate times the density of moist air. The density is found
6
using the psychrometric chart and taking the inverse of the humid volume that
corresponds to the dry bulb temperature and relative humidity measured.
  V *   A * v * 
m
(3)
Though the mass flow rate calculated is found at the exhaust, this value can also
be applied to the intake as well. Equation (4) demonstrates a mass balance for the system,
showing that mass is neither created nor destroyed.
`
 in  m
 exit
m
(4)
The rate of heat transfer can now be determined through the First Law of
Thermodynamics, also known as the energy equation, for an open system. This equation
is demonstrated below in equation (5).
 
 
 in hˆin  m
 exit hˆexit  Qin  0
m
(5)
When comparing the solar radiation receiving heat into the building, theoretically
the heat absorbed by the air flow should be greater. If the solar radiation were greater,
then customers and staff would complain about the air-conditioning not functioning
properly.
7
EQUIPMENT
To perform this experiment, many devices were required. Due to the height of the
cooling tower and location of the exhaust fans, a ladder was necessary. An anemometer
(#0 in the lab 119 room) was used to measure both temperature and velocity. To ensure
that each point was measured accurately at the same distance from the center for each
fan, a yard stick was used. A sling psychrometer (the square black one in lab 119 room)
was used to measure wet and dry bulb temperatures at the inlet. Figure 3, below, shows
pictures of the equipment used.
Figure 3: Picture of Equipment Used
For calculation of absolute humidity and enthalpy, a psychrometric chart was used.
8
PROCEDURE
First, the ambient air temperature was recorded on site. Then, after looking at the
four exhaust fans from operating system, A, evaluation of which points would be best to
take samples from was discussed. Velocity and temperature were measured and recorded
at the inner, middle, and edge of each fan; however, the same radial measurements were
not taken for both weeks 1 and 2. Figure 4, below, shows a schematic of the exhaust fan
and locations of where points were measured from during both experiments. Week 1
points are located on the left of the schematic in red, while week 2 points are located on
the right in purple. The center cap was measured to be 0.14 ft and the radius across the
fan was measured to be 1.30 ft.
Week 2
Week 1
0.22 ft
0.55 ft
0.72 ft
0.97 ft
1.2 ft
1.25 ft
Figure 4: Schematic of Exhaust Fan with Measured Locations
Velocity, measured in feet per minute, from week 1 and week 2 across all four
fans is shown below. The first measurement was taken at the north-most point of the fan
and last measurement at the south-most point.
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Measurement
1
2
3
4
5
Measurement
1
2
3
4
5
6
Week #1
Fan #1
Fan #2
Radial Position
Velocity
Velocity
(ft)
(ft/min)
(ft/min)
1.25
1908
2066
0.97
2206
2046
0.55
330
511
0.97
1400
1338
1.25
1537
1574
Table 1: Measured Velocities from Week 1
Fan #3
Velocity
(ft/min)
2106
2086
452
1712
1771
Fan #4
Velocity
(ft/min)
1751
2125
826
1672
2223
Week #2
Fan #1
Fan #2
Radial Position
Velocity
Velocity
(ft)
(ft/min)
(ft/min)
1.22
2105
2381
0.72
2027
2164
0.22
196
334
0.22
728
846
0.72
2223
2243
1.22
2027
2145
Table 2: Measured Velocities from Week 2
Fan #3
Velocity
(ft/min)
2066
2223
354
688
2105
1849
Fan #4
Velocity
(ft/min)
2223
2145
177
806
2164
2086
Temperature, measured simultaneously with velocity through use of the
anemometer, is shown below for both experiments. The ambient air temperature during
week 1 was 99.5ºF and during week 2 was 97.6ºF.
Week #1 (Tambient = 99.5 ºF)
Fan #1
Fan #2
Measurement
1
2
3
4
5
Measurement
Fan #3
Radial Position
(ft)
Temp (ºF)
Temp (ºF)
Temp (ºF)
1.25
103.5
105.8
108.0
0.97
106.7
104.8
108.6
0.55
109.8
106.2
107.7
0.97
110.0
107.2
108.8
1.25
106.5
105.9
107.5
Table 3: Recorded Temperatures for Week 1
Week #2 (Tambient = 97.6 ºF)
Fan #1
Fan #2
Radial Position
Temp (ºF)
Temp (ºF)
10
Fan #3
Temp (ºF)
Fan #4
Temp (ºF)
107.7
110.4
110.9
110.7
111.3
Fan #4
Temp (ºF)
(ft)
1.22
102.4
102.3
99.6
0.72
106.8
106.1
100.8
0.22
105.5
106.6
102.3
0.22
105.5
106.7
101.7
0.72
107.4
106.5
102.2
1.22
103.2
105.4
100
Table 4: Measured Temperatures for Week 2
1
2
3
4
5
6
97.7
100.4
103.5
103.3
102.7
102.2
For week 2, dry bulb temperatures were then measured and recorded across the
eight different intakes. Wet bulb temperatures and percent relative humidity were
measured from the outside intake. It is important to note that dry bulb temperatures were
measured with an anemometer while wet bulb temperatures were measured with a
psychrometer. The data measured is shown below.
Week #2 Air Intake Temperatures
Fan #1
Fan #2
Fan #3
Fan #4
Tinside intake(ºF) Toutside intake(ºF) T (wet bulb) (°F) Relative Humidity (%)
95.7
94.5
75.2
40
93.6
95.3
74.8
39
91.2
93.1
74.8
43
90.1
90.5
73.4
46
Table 5: Recorded Wet Bulb and Dry Bulb Temperatures at Air Intake
Once completed, compressor run times were recorded from the user’s guide over
a five days period in order to compare systems A and B. The user’s guide is located on
the west side of the cooling tower. Recorded running hours are shown in Table 6. The
values are expressed in terms of hours per day in the results section.
Date
8/19/2008
8/20/2008
8/21/2008
8/22/2008
8/25/2008
Starts
1913
1915
1917
1919
1923
Compressor A
Compressor B
Running Hours
Starts
Running Hours
3017
1913
2885
3039
1915
2887
3060
1917
2890
3081
1919
2891
3125
1923
2920
Table 6: Compressor Run Times
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RESULTS
The temperatures across each radial position and fan were averaged and then
graphed for comparative purposes, as demonstrated below for both experiments.
112
110
Temperature (F)
108
106
ΔT ≈ 10 ºF
Fan #1
Fan #2
Fan #3
Fan #4
Ambient Inlet
104
102
100
98
96
0.0
0.2
0.4
0.6
0.8
1.0
Radial Position (ft)
Figure 5: Temperature Dependence for Week 1
12
1.2
1.4
112
110
Fan#1
Fan#2
Fan#3
Fan#4
Ambient Inlet
Temperature (F)
108
106
104
ΔT ≈ 6
ºF
102
100
98
96
0
0.2
0.4
0.6
0.8
1
1.2
1.4
Radial position (ft)
Figure 6: Temperature Dependence for Week 2
Since velocity changes throughout different radii on each fan, the midpoints
between each radius measured were averaged together and assumed to be the same for
each band. Below, in Figure 7, values “r1” through “r4” labeled in red were the measured
distances. Values “R2” and “R3” labeled in blue and represented by dashed lines, were
the midpoint averages while value “R1” was the radius of the center cap and value “R4”
was the outer fan radius. Equation (6) shown below demonstrates the calculation for the
area in a band.
A1 = 2 Π * (R2² - R1²)
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(6)
Figure 7: Schematic of Fan to Determine Area Calculations
Table 6, below, shows the area and average velocity at each radii across each fan.
Radial
Position (ft)
0.22
0.72
1.22
Fan #1
Fan #2
Fan #3
Area
Velocity
Velocity
Velocity
(ft²)
(ft/min)
(ft/min)
(ft/min)
0.63
462
590
521
2.26
2125
2204
2164
2.38
2066
2263
1958
Table 6: Area and Average Velocity for Week 2
Fan #4
Velocity
(ft/min)
492
2155
2154
Through use of the psychrometric chart, average humid volume of the moist air
across the entire unit averaged 14.6 ft³/lb; therefore the average density of the moist air is
0.069 lb/ft³. With this information, mass flow rate is easily attained and expressed below
in Figure 8.
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50000
45000
Mass Flow Rate (lb/hr)
40000
35000
30000
25000
20000
15000
10000
5000
0
Fan 1= 41,200 Fan 2= 44,200 Fan 3= 40,600 Fan 4= 42,400
Figure 8: Mass Flow Rate of Each Fan for Week 2
The enthalpy was then solved for two different ways by use of the psychrometric
chart. At the intake, the relative humidity was measured and used with the dry bulb
temperature to solve for enthalpy in. With those same values, the absolute humidity was
then obtained to assist in solving for the exit enthalpy. Though the relative humidity will
differ at the intake and exhaust, the absolute humidity will remain the same. This value
was then used with the dry bulb temperature at the exhaust to solve for the exit enthalpy.
These values are best expressed in Figure 9, in terms of Btu per pound.
15
36
34
34
Enthalpy (Btu/lb)
34
33
32
32
31
31
31
30
30
Inlet
28
Exhaust
26
1
2
Fan #
3
4
Figure 9: Enthalpy at the Intake and Exhaust for Each Fan
With the enthalpies attained, the heat into the system is solved for by
manipulation of the energy equation, shown below in Equation (7).
Q in  m * (hˆexit  hˆin )
(7)
The results for energy in, or heat absorbed by the air, are graphically expressed
below in Figure 10. The sum of the heat absorbed by all of the fans is 390,000 ± 9,500
Btu/hour.
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120000
100000
Heat (Btu/hr)
80000
60000
40000
20000
0
Fan 1 =103,000
Fan 2 =105,000
Fan 3 =75,500
Fan 4 =106,000
Total Heat Absorbed= 390,000
Figure 10: Heat Absorbed by Air for Each Fan
Now that the heat absorbed has been determined, it can be compared to the solar
radiation cast onto the building. The direct solar radiation for September 2, the day of the
second experiment, is approximated to be 600 W/m² around 3 pm. For calculation
purposes solar radiation was converted to 190 Btu/hr/ft². Given the building size of
23,000 square feet calculations, as shown in the appendix, result in 4 million Btu/hour of
radiation onto the building during the second experiment in September [2].
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Figure 11: ORNL Data for Solar Radiation on September 2, 2008
The data recorded for the compressor run times enabled the running hours per day
to be determined. To find the unit running hours, the difference was taken between two
days. This is the reason days one and five are not represented on the bar graph shown
below, in Figure 11.
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Unit Running Hours
24
20
Unit A
16
Unit B
12
8
4
0
2
3
Days
Figure 12: Running Hours per Unit per Day
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4
DISCUSSION OF RESULTS
It is observed while comparing temperature dependence in Figures 5 and 6 that
week 1 showed higher temperatures from fans 1 and 4, while week 2 showed higher
temperatures from fans 1 and 2. Week one showed the lowest temperatures at fan 2 while
week 2 showed the lowest temperatures at fan 3.
Looking back at Figure 8, it is observed that the mass flow rates across each fan
minimally varied. They ranged from 40,000-44,000 lb/hr.
The heat absorbed into the system, depicted in Figure 10, shows similar heat for
fans 1, 2, and 4; however fan 3 absorbs less heat by comparison. This is due to a
combination of a lower mass flow rate and smaller change in enthalpy, in comparison to
fans 1 and 2. Fan 4 has the same change in enthalpy, however, does not result in less heat
absorbed due to the higher mass flow rate. It is suspected that velocity measurements
were recorded inaccurately, the day of the experiment.
The amount of solar radiation brought onto the building does not take into
account insulation, reflective properties within the roofing material, or the size and
amount of windows throughout the building. When comparing this value to the heat
absorbed by the air, compressor B, which also runs daily, is unaccounted for.
When comparing units A and B, it is observed the unit A consistently runs for a
much longer period throughout the day than unit B. On day 4, the units did not run a full
24 hour period, which means at some point the outdoor temperature reached the desired
indoor temperature for that day.
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CONCLUSIONS
Through analysis many observations about the behavior of the Challenger Center
cooling tower can be made, but only a few conclusions are determined. It is concluded
that unit A is very close to meeting the demands for the building. With regards to
temperature, while performing the experiments it was thought that the sunlight and
shading of the units affected the temperature at the exit. Based on the results, that theory
was proven incorrect. When comparing the velocity from fan 3 during weeks 1 and 2, it is
concluded that the data was measured or recorded inaccurately and produced a lower heat
value in comparison to fans 1, 2 and 4. However, this is not proven since relative
humidity readings were taken during week 1.
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RECOMMENDATIONS
When analyzing the energy of a system through measurements, it is
recommended, for comparative purposes, that the measurements be taken at the exact
same points. It is difficult to compare velocities when there is much variance throughout
the different radii of the fans. The next recommendation is to check and compare all
equipment for function and calibration before depending on it for your results. Also,
because there are different instruments that look similar make sure to record or mark
which one is being used during the experiment. This will ensure that, over multiple
experiments, the same instrument is being used.
The approach for determining the amount of solar radiation received by the
building is suspected to be inaccurate, due to the result in such a high heat value. An
accurate approach has not been solved for, but it is recommended that the sizing of the
windows be taken into account and heat transfer between the glass in the future.
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REFERENCES
1. Felder, Richard M. and Rousseau, Ronald W. Elementary Principles of Chemical
Processes, 3rd edition. John Wiley and Sons, 2000.
2. http://www.nrel.gov
ORNL RSR Data Plot for September 2, 2008; October 7, 2008.
3. http://chem.engr.utc.edu/engr435/ChallengerCenter/cooling-tower.htm
Smith, Jennifer; Staton, Craig, “Challenger Center Cooling Tower Analysis”,
November 30, 2000.
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