Challenger Center Cooling Tower

Transcription

1 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

2 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 F at the exhaust fans, and 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

3 TABLE OF CONTENTS Introduction Theory Equipment Procedure Results Discussion of Results Conclusions Recommendations References Appendix

4 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 Outside Intake B Inside Intake Outside Intake A Inner Intake B 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

5 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

6 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* (2) v avg 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

7 using the psychrometric chart and taking the inverse of the humid volume that corresponds to the dry bulb temperature and relative humidity measured. m & = V & * ρ = A* v * ρ (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. ` m & = & (4) in m exit 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 ( h ˆ ) m& ( hˆ ) + Q = 0 m & (5) in exit exit in 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

8 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

9 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. 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 1 Week ft 0.97 ft 1.25 ft 0.22 ft 0.72 ft 1.2 ft Figure 4: Schematic of Exhaust Fan with Measured Locations, 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. 9

10 Week #1 Fan #1 Fan #2 Fan #3 Fan #4 Measurement Radial Position (ft) Table 1: Measured Velocities from Week 1 Week #2 Fan #1 Fan #2 Fan #3 Fan #4 Measurement Radial Position (ft) Table 2: Measured Velocities from Week 2 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. Measurement Week #1 (T ambient = 99.5 ºF) Fan #1 Fan #2 Fan #3 Fan #4 Radial Position (ft) Temp (ºF) Temp (ºF) Temp (ºF) Temp (ºF) Table 3: Recorded Temperatures for Week 1 Week #2 (T ambient = 97.6 ºF) Fan #1 Fan #2 Fan #3 Fan #4 Measurement Radial Position Temp (ºF) Temp (ºF) Temp (ºF) Temp (ºF) 10

11 (ft) Table 4: Measured Temperatures for Week 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 (%) 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. Compressor A Compressor B Date Starts Running Hours Starts Running Hours 8/19/ /20/ /21/ /22/ /25/ Table 6: Compressor Run Times 11

12 RESULTS The temperatures across each radial position and fan were averaged and then graphed for comparative purposes, as demonstrated below for both experiments Temperature (F) ΔT 10 ºF Fan #1 Fan #2 Fan #3 Fan #4 Ambient Inlet Radial Position (ft) Figure 5: Temperature Dependence for Week 1 12

13 112 Temperature (F) ΔT 6 Fan#1 Fan#2 Fan#3 Fan#4 Ambient Inlet 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) (6) 13

14 Figure 7: Schematic of Fan to Determine Area Calculations Table 6, below, shows the area and average velocity at each radii across each fan. Fan #1 Fan #2 Fan #3 Fan #4 Radial Position (ft) Area (ft²) Table 6: Area and Average for Week 2 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 lb/ft³. With this information, mass flow rate is easily attained and expressed below in Figure 8. 14

15 Mass Flow Rate (lb/hr) 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

16 36 Enthalpy (Btu/lb) Inlet Exhaust Fan # 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 & = m& ( hˆ hˆ ) (7) in * exit in 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. 16

17 Heat (Btu/hr) 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]. 17

18 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

19 Unit Running Hours Unit A Unit B Days Figure 12: Running Hours per Unit per Day 19

20 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. 20

21 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. 21

22 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. 22

23 REFERENCES 1. Felder, Richard M. and Rousseau, Ronald W. Elementary Principles of Chemical Processes, 3 rd edition. John Wiley and Sons, ORNL RSR Data Plot for September 2, 2008; October 7, Smith, Jennifer; Staton, Craig, Challenger Center Cooling Tower Analysis, November 30,