The Evaluation and Design of the Ventilation System Within Mansfield Dam (Lake Travis)

Transcription

1 The Evaluation and Design of the Ventilation System Within Mansfield Dam (Lake Travis) Submitted to: Andy Sumner, P.E., Senior Engineer, Dam and Hydroelectric Division Lower Colorado River Authority P.O. Box 8, Buchanan Dam, Texas Prepared by: Clark Hughes, Team Leader Stephen Johnson Matthew Payne Mechanical Engineering Design Projects Program The University of Texas at Austin Austin, TX Summer 1997

2 Acknowledgments The design team would like to thank many people for their contributions to this project. The team would like to thank the Lower Colorado River Authority for sponsoring this design project. Mark Johnson supplied the team with valuable environmental information, and Kim Conley provided historical background on LCRA and Mansfield Dam. Andy Sumner and Bob Fohn served as the team s main contacts at LCRA. They provided access to the dam and answered many of the team s questions about the project. The team would also like to thank Dr. Gary Vliet of The University of Texas Mechanical Engineering Department. Dr. Vliet served as the team s faculty advisor and spent many hours with the team in weekly meetings. He provided assistance with evaluating heat exchangers, modeling the dam, and sizing the conditioning system. The faculty and staff of the Mechanical Engineering Design Projects Program have also been a great help. Dr. Steven Nichols lectures on professionalism in the field of engineering were very insightful and educational. Dr. Thomas Krueger assisted the design team in preparing graphics for presentations and papers. Teaching assistant Jeff Norrell s advice and attention to detail helped the team produce a thorough report. The team would also like to thank Sarah Searcy and Dee Davenport for running the Design Projects Program on a daily basis.

3 Executive Summary The equipment within Mansfield Dam, particularly in tunnels housing floodgate equipment, is being corroded by moisture that condenses out of the air. Currently, air enters the dam through passive ventilation. This air carries moisture which condenses onto cool surfaces within the dam. This condensation is particularly apparent in the summer months when the temperature difference between cool lake water and warm ambient air is high. Maintenance to deal with the effects of corrosion currently costs the Lower Colorado River Authority (LCRA) $65,000 a year. In addition to corrosion problems, employee safety is threatened by slippery floors, walls, and stairs. These problems can be solved by removing moisture from the air before the air enters the dam. Shortly after the construction of Mansfield Dam in the 1940's, dehumidification equipment was installed in a concrete bunker at one end of the dam. This equipment ceased to operate in the 1960's. The design team has examined the original system and other potential dehumidification solutions. The team has found that a rehabilitated system will not prevent condensation on the coolest surfaces within the gate galleries. Other dehumidification systems such as desiccant and vapor-compression systems have also been examined. The team recommends that a water cooled vapor-compression dehumidification system be installed in place of the old equipment. Such a system offers potential savings of $582,000 in maintenance costs over a twenty year life. In addition to eliminating corrosion and safety hazards, this system will provide cooling air to the power house. A more detailed discussion of the humidity problem and method used to find a dehumidification solution follows in the rest of this report. In addition to the team s final recommendation, the portions of this report that may be most beneficial to LCRA are the dehumidification specification sheet, evaporation model, heat transfer model, and air flow model. The design team developed the models in order to specify the system and developed the specification sheet to consolidate dehumidification system requirements. ix

4 I. Introduction The Lower Colorado River Authority (LCRA) sponsored a team from The University of Texas at Austin to recommend a solution for corrosion and safety problems associated with high humidity levels inside Mansfield Dam. The design team consists of three mechanical engineering students from the Design Projects Program. These students are Clark Hughes (Team Leader), Stephen Johnson, and Matthew Payne. Clark Hughes has completed a full co-op term designing and troubleshooting electronic systems at Applied Research Laboratories. Stephen Johnson interned at Intermedics Orthopedics, an Austin-based manufacturer of artificial knees, hips, and shoulders. He also constructed a prototype weigh-in-motion system for heavy vehicles as a research assistant at Penn State University. Matthew Payne worked at the Center for Electromechanics in Austin designing tooling for an electromagnetic cannon. He has also conducted prototype assembly and testing of this cannon. Mansfield Dam contains 5.6 km (3.5 mi) of tunnels which house floodgate machinery and other equipment with exposed metal surfaces, such as metal carts, pulleys, and chains. A tunnel cross-section is shown in Figure 1. The tunnels in the dam have 23 cm to 30 cm (9 in to 12 in) wide drainage 1.0 m 2.0 m 3.43 m trenches that contain drainage and dam seepage water. Water from moisture laden air Drainage trench Figure 1. Tunnel cross section.

5 condenses on cool equipment inside the dam. The resulting surface corrosion makes repainting and reservicing the floodgate machinery necessary on an annual basis [Fohn, 6/11/97]. Additionally, the moisture causes electrical shorts and corrosion of electrical equipment. Water also condenses out of the air on the floors and working surfaces, creating a slippery, hazardous working environment. Figure 2 illustrates condensation on tunnel walls. Figure 2. Condensation on tunnel wall. LCRA asked the design team to examine dehumidification systems including, but not restricted to, a rehabilitated version of a system that was originally installed in the dam. The team determined that the existing equipment cannot be used to completely eliminate condensation within the dam. The original dehumidification equipment utilized cool lake water to dehumidify air entering the dam, but the inner surfaces of the dam and floodgate equipment are also in contact with the same water. Therefore, the dew point of the air cannot be lowered below the temperature of these cool surfaces. However, a 2

6 number of alternative solutions to dehumidify the air in the tunnels exist. These solutions include vapor-compression systems, desiccant systems, and a water cooled heat exchanger similar to the existing dehumidification equipment. This paper presents the design team s method used to solve this dehumidification problem in a systematic manner and outlines some concept variants based on the results of research and system models. The report concludes with the recommendation of installing an air cooled vapor-compression system to reduce the humidity levels in Mansfield Dam. II. Problem Statement Moisture laden air in Mansfield Dam causes equipment corrosion, safety problems, and uncomfortable working conditions. As a first step in solving the problems associated with moisture in the dam, the design team gathered historical information about LCRA and Mansfield Dam. A summary of this research and an investigation of the problems in the dam follow. 2.1 Background on LCRA and Mansfield Dam The Texas Legislature created the Lower Colorado River Authority (LCRA) in 1934 as a conservation and reclamation district to improve the life of Texans living in the Hill Country and along the central and southern Colorado River basin [Williams, 1985]. 3

7 The idea to dam the Colorado River for electrical generation, water supply, and flood control existed as early as 1854, but support for the dams did not materialize until the 1920s [Williams, 1985]. Around the turn of the century, the City of Austin built several dams, but these structures were severely damaged or destroyed by flooding of the Colorado River. Between 1900 and 1915 alone, the Colorado River flooded 20 times killing 94 people and destroying millions of dollars worth of property [Adams, 1990]. After four legislative sessions, lengthy fights with west Texas landowners and private utility companies, and numerous concessions to all involved parties, the Texas Legislature created LCRA on November 10, 1934 [Williams, 1985]. Six months earlier President Roosevelt had authorized the release of Public Works Administration (PWA) funds to the State of Texas, but the loans had to be handled by a public agency. LCRA became this agency. Through LCRA management of PWA funds, four dams were built between 1935 and These dams were Buchanan Dam, Inks Dam, Mansfield Dam, and Tom Miller Dam [Williams, 1985]. The construction of Mansfield Dam, originally Marshall Ford Dam, began in February of 1937 and ended in May The dam is m (7098 ft) long, 81.1 m (266 ft) tall, 9.1 m (30 ft) wide at the top, and 64.9 m (213 ft) wide at the bottom. Figure 3 presents a dimensioned sketch of Mansfield Dam. Although equipped with 24 floodgates, only six have ever been opened at one time [LCRA Chronicle, 1991]. These floodgates reside within two main galleries, or tunnels, inside the dam. In addition to the two larger gallery spaces, there are four other smaller tunnels within the dam used for drainage and inspection purposes. 4

8 9.1 m (30 ft) Gate Galleries 81.1 m (266 ft) 64.9 m (213 ft) Dehumidification Bunker 2161 m (7089 ft) Figure 3. Schematic of Mansfield Dam with dimensions. 2.2 Problems Associated with Moisture Figure 4 presents a cross-sectional view of Mansfield Dam. The dam contains a network of 5.6 kilometers (3.5 miles) of tunnels that provides access to floodgate machinery. Condensation in the dam leads to corrosion of mechanical and electrical systems and presents a safety hazard. The floodgate machinery requires annual repainting and maintenance. This maintenance costs LCRA $65,000 a year [Sumner, 7/9/97]. The floodgate machinery is cool because it contacts the water of Lake Travis. The water at the level of the floodgates, about 45.7 m (150 ft) below the surface of Lake Travis, is between 10 C (50 F) and 12.8 C (55 F) throughout the year [Limnographical, 1965]. 5

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10 Electrical systems are also affected by the moisture in the dam. Although wiring in the dam is contained in waterproof conduits, water has been found in sealed electrical boxes [Fohn, 6/11/97]. Such conditions significantly reduce the operational life of the electrical systems, increase maintenance costs, and pose the threat of electrical shorts. Safety is also a concern. The moisture in the dam makes floors, walls, and stairs slippery and increases the likelihood of a fall. There is also the possibility of electrocution from wet wires. Furthermore, the damp environment makes the dam an attractive home for pests such as spiders and coral snakes [Fohn, 6/11/97]. 2.3 Existing Dehumidification System In an effort to prevent condensation inside the dam, LCRA installed a dehumidification system shortly after the construction of Mansfield Dam. A schematic of this existing system is shown in Figure 5. warm moist air cool lake water filter condensate heat exchanger heat added if necessary water discharge blower cool, drier air warmed lake water Figure 5. Original dehumidification system. 7

11 This dehumidification system was designed to utilize the low temperature of the lake water to dry the air. Cool water was routed from Lake Travis through one of the floodgates and fed into a heat exchanger. Ambient air was drawn across this heat exchanger, and the air was cooled by lake water. As the air was cooled to a temperature below its dew point, condensate formed on the heat exchanger coils and was routed out of the system. The air was then both drier and cooler than it was in its original state. If the air was too cool for employee comfort, heat was added by means of a resistive heating network before the air was blown into the dam tunnels. Currently, the existing dehumidification equipment in Mansfield Dam is inoperable due to failure of the blower, damage to the heat exchanger, and missing parts [Fohn, 6/11/97]. The design team interviewed two LCRA personnel to obtain information about the original dehumidification in the dam. Bob Fohn, the Area Supervisor for Austin and Mansfield Dams, has been associated with Mansfield Dam for 30 years. He remembers the dehumidification system working through the first year of his employment, around The design team also met with Don Smith, a former LCRA employee who worked with the original dehumidification system. Problems with the existing system began with the blower. LCRA believes that the blower began vibrating violently because the blower went out of balance. LCRA sent the blower off for balancing, but the problem was not corrected. This violent vibration has prevented operation of the system since that time [Fohn, 6/11/97]. An interview with Bob Duncan of the Tisdale Corporation revealed that blowers and fans are commonly misdiagnosed as having gone out of balance when, in fact, the machinery is run at the 8

12 wrong speed for the operating pressure differential [Duncan, 7/17/97]. LCRA may test the blower machinery in the future to evaluate possible rehabilitation. Time and neglect have taken their tolls as well. Practically every component of the existing dehumidification system is in need of replacement or repair, including the filters, the cold water supply line, the heat exchanger, the blower, and the heating elements [Sumner, 6/11/97]. These components have experienced 20 to 30 years of corrosion and lack of maintenance. Some heating elements and sections of the cold water supply line are missing. 2.4 Project Requirements The primary objective of the design team's analysis was to reduce problems associated with humidity. LCRA wants to prolong the life of equipment such as floodgates, pumps, and other electrical and mechanical equipment inside the tunnels within Mansfield Dam. Interviews with dam personnel and first-hand observation revealed that moisture from high humidity in the tunnels had caused surface corrosion on mechanical and electrical equipment, rust on exposed metal, and slippery conditions in some locations [Fohn, 6/11/97]. As a first step toward finding a solution, the team spoke with a number of employees at Mansfield Dam and then generated a list of customer needs based on this contact. The needs address problems associated with maintenance as well as safety concerns. Table 2 outlines these customer needs and also presents a number of different engineering principles that address these needs. The intersection of a need and a principle 9

13 that would meet that particular need is marked by an 'X' in the table. As indicated in Table 1, each customer need can be addressed independently. For example, the customer need of reducing the danger of slipping can be met by increasing the coefficient of friction on the tunnel floors. The coefficient of friction can be directly increased by installing metal grating and handrails in the tunnels, but the coefficient of friction can also be indirectly increased by reducing humidity within the dam. Reducing humidity meets all of the customer needs. On this basis, the design team decided to focus on the problem of reducing the humidity inside the dam. Table 1. Customer Needs and Engineering Principles. Engineering principles Customer needs Reduce rusting of exposed metal Reduce danger of slipping Reduce repainting of equipment Increase equipment reliability Reduce electrical hazards Increase comfort Reduce surface area of rust, cm 2 Increase coefficient of friction within tunnels, µ Increase distance that separates surfaces of concern from moisture, mm Install mechanical equipment that has an increased lifetime, t Install electrical equipment that has an increased lifetime, t Reduce absolute humidity within the dam, ω X X X X X X X X X X X X X 10

14 2.5 Constraints The design team identified some constraints that warranted consideration. The original dehumidification system utilized the tunnels in the dam to distribute treated air. The design team decided to use these tunnels for air distribution within the dam because space for ductwork in the tunnels is extremely limited. Also, free space for dehumidification equipment within the tunnels does not exist without modifying existing equipment placement in the tunnels. This means that the system that the design team recommends must be able to fit in the existing dehumidification bunker. Another constraint is the hot and humid Texas environment. Many parameters including dry bulb temperature, wet bulb temperature, and absolute humidity can be used to define this constraint. Dry bulb temperature is a temperature measurement of ambient air that does not take humidity into account. Wet bulb temperature does take evaporation into account. Wet bulb temperature is taken by using a thermometer that has water evaporating off of the bulb. These two temperatures, wet and dry bulb, are used together to determine absolute humidity. Absolute humidity is a dimensionless ratio that relates the mass of moisture in the air to the mass of the dry air. Figure 6 shows a psychrometric chart used to relate wet bulb temperature, dry bulb temperature, and absolute humidity. Relative humidity, also found on the psychrometric chart, gives a percentage of moisture contained in the air relative to how much water the air can hold when saturated. When the ratio of water vapor to air remains constant, an increase in dry bulb temperature lowers the relative humidity. This means that more moisture can be absorbed by the air. 11

15 100% Relative Humidity 50% Relative Humidity Increasing Dry Bulb Temperature Increasing Wet Bulb Temperature Increasing Absolute Humidity Figure 6. Psychrometric chart. The team had to specify system suitable for operation in Central Texas. The American Society of Heating, Refrigerating and Air Conditioning Engineers (ASHRAE) design dry bulb temperature for Austin, Texas is 38 C (100 F), and the design wet bulb temperature is 25.6 C (100 F). These design conditions give an absolute humidity of kg water /kg dry air [ASHRAE, 1989]. These design conditions represent temperatures and humidities that are exceeded only one day out of every one hundred. The dehumidification system will have to dry the air while maintaining ASHRAE standards for human comfort. This means that the treated air must have an absolute humidity within the bounds of kg water /kg dry air to kg water /kg dry air. The treated air must also have a dry bulb temperature between 19 C (66 F) and 27 C (81 F) [ASHRAE, 1989]. 12

16 The final project constraint is that the system cannot interfere with present, normal dam operation. This constraint requires the design team to recommend an unobtrusive system that does not change any of the activities and operations that regularly take place in the dam. For example, the team cannot recommend a system that effects floodgate operation. 2.6 Criteria For Evaluation The design team determined that reducing humidity in the dam tunnels is the main objective of this project. Therefore, a numerical value that describes the moisture content in the air and on equipment in the tunnels is used to evaluate the effectiveness of the team's final design. Specifically, the design team hopes to eliminate all condensation on interior surfaces of the dam. Even a small amount of condensation on a metal surface is sufficient to cause corrosion. By definition, wet corrosion occurs when liquids are present on exposed surfaces [Bosich, 1970]. There are four components that are necessary for the corrosion of metals: an anode, a cathode, an electrolyte, and an electrical circuit. The condensation, even a small amount, acts as an electrolyte. This report specifies a system that prevents corrosion by removing the electrolyte [Barton, 1976]. In order to prevent condensation on the tunnel floors, walls, and floodgate equipment surfaces, the dehumidification system must lower the dew point of the air below the temperature of these surfaces [Clifford, 1990]. Achieving this dew point prevents water from condensing on these surfaces. Heating the surfaces to raise surface 13

17 dew point is another possible method of locally preventing condensation, but the team did not pursue this principle. Heating these surfaces would be unfeasible because they are in direct contact with the cool water of Lake Travis. The next section of this report outlines the methodology the design team used to develop possible solutions to the problem. III. Methodology This section of the proposal begins with a presentation of the design team's initial analysis of the engineering tasks inherent in this problem. The section continues with the methodology that the team utilized to arrive at a solution to this problem. 3.1 Preliminary Considerations As an initial analysis, the design team examined the moisture content of the air that cycles through the dam. Figure 7 presents a schematic of this cycle. Warm moist air enters the dam from the outside environment. This air passes through a conditioning system that removes moisture. Then the air enters the dam. As this dry air flows through the tunnels, it absorbs moisture from drainage trenches, wetted wall surfaces, and damp machinery. Finally, the air returns to the surrounding environment. Once the wall surfaces and machinery dry out, the drainage trenches still supply moisture to the air because they constantly collect seepage throughout the dam [Sumner, 6/11/97]. 14

18 Moist air 1. Outdoors Moist air 2. Conditioning Equipment Dry air 3. Dam Tunnels Water from air Evaporated water from trenches and walls Figure 7. Sources of moisture in air. Based on this initial analysis, the team identified several parameters that must be quantified in order to specify the performance of the dehumidification system to be designed. Figure 8 shows these parameters and illustrates the order that they were applied in order to arrive at a satisfactory set of dehumidification system specifications. Evaporation from Drainage Trenches and Walls, &m (kg/s) Air Flow Pattern through Dam, Q (m 3 /s) Temp. of Inner Surfaces, T ( C) State of Outside Air (temp., humidity), T ( C), ω Conditions of Treated Air, &m, Q, T Temp. of Water Entering the System, T ( C) Variables &m - mass flow rate Q- volumetric flow rate T- temperature ω- absolute humidity Specification of Conditioning System, ω, Q, T Figure 8. Path to system specification. 15

19 3.1.1 Temperature of Inner Surfaces The team identified the temperature of the coolest surfaces in the dam as a key factor in determining the necessary state of the air exiting the dehumidification system. To eliminate condensation on floodgate equipment and tunnel surfaces, it is necessary to lower the dew point of air in the dam to a temperature lower than that of those cool surfaces. The design team used two approaches to determine the temperature of these surfaces. First, the design team acquired tail race temperature data for the time period July 1995 to July 1997 [Sumner, 7/17/97]. The tail race of the dam is the body of water directly downstream of the power generation equipment, so the temperature of this water should coincide with the temperature of the lake water as it enters the dam on the upstream side. The design team met with James Piper, the lead diver for Applied Research Laboratories (ARL), on June 20, 1997 and Dr. Gary Vliet of the University of Texas at Austin on June 11, Both of these men confirmed that the tail race temperatures should be the same as the temperatures of the water entering the dam on the upstream side. This tail race temperature data is provided in Appendix A. An average of this data over the months of July and August yields a water temperature of 22.5 C (72.5 F) [Sumner, 7/17/97]. This value seems high, especially when compared to some of the design team s temperature measurements. During a trip to the dam on June 23, 1997, the team measured the temperature of water exiting a floodgate through a small valve as 18.3 C (64.9 F) using a hand-held digital thermometer. This contradicts the tail race data, 16

20 which indicates warmer water temperatures. In order to obtain more plausible temperature data, the team then turned to a computational approach. The design team located a record of temperature data for Lake Travis compiled by the Defense Research Laboratories, now called Applied Research Laboratories [Limnographical, 1965]. This data provides water temperatures at various depths down to 150 feet in Lake Travis from 1963 through The design team developed a Matlab routine that incorporates this water temperature data to generate a thermal model of the cross-section of Mansfield Dam. The results from this model are summarized in Figure 9, which shows the temperatures at all locations within the dam cross-section Distance (meters) Figure 9. Temperature profile of dam cross-section. 17

21 The Matlab script is called tdam.m and is presented in Appendix B. The water temperature data is incorporated by the program to calculate the temperatures along the upstream side of the dam. The exposed surfaces of the dam are assumed to be 29.5 C (85 F) during the months of July and August, and the bottom surface is assumed to be the same temperature as the bottom of the lake. Given these boundary conditions, the program calculates the temperatures at all locations within the cross-section and displays this temperature data in a color-coded plot. From this plot, the design team determined the temperature of the surfaces inside the dam. The surfaces are indicated as tunnels in Figures 4 and 9. This computational approach is only as accurate as the assumptions upon which it is based. Errors in any of the boundary temperatures discussed above will adversely affect the entire model. Although this model is clearly sensitive to errors, the design team believes that this approach provides more accurate data than the tail race temperature records. This thermal model indicates that the surface wall temperature in the upstream floodgate gallery is approximately 15.5 C (60 F). This value is lower than the value suggested by the tail race data. The team believes that the tail race data provides artificially high temperature data for two reasons. First, the tail race is exposed to warm ambient conditions and is heated to temperatures above those of the water entering the upstream side of the dam. Second, the generators are not always in operation, which means that the tail race water is somewhat stagnant. This factor magnifies the effect of ambient conditions on the temperature of this body of water. These two factors help explain why the tail race average temperature is higher than the 18

22 temperature predicted by the thermal model. Regardless of the relative accuracy of these two sources of temperature data, the design team feels that the thermal model temperatures should be used to represent the surface temperatures simply due to the fact that they are lower and represent worst case conditions. Following this reasoning, the design team used 15.5 C (60 F), the value indicated by the thermal model, as the worst-case surface temperature for the upstream gate gallery Air Flow Model An understanding of the air flow patterns in the dam allows the design team to identify locations that will be the last to receive conditioned air. As dry air flows through tunnels in the dam, it absorbs moisture from wetted surfaces. Air reaching locations furthest from the dehumidification system must still be dry enough to prevent condensation on cool surfaces. The team had to design a system that would provide sufficiently dry air to these locations. The team s first step in determining the air flow pattern through the dam was to inspect the tunnels in the dam. There are six tunnels, also known as galleries, in the dam with a total length of approximately 5.6 km (3.5 mi). The six galleries are connected to one another by vertical ventilation shafts [LCRA Blueprint 249-D-833, 1939]. Most of these shafts were boarded shut when the original dehumidification system failed [Fohn, 6/11/97]. To estimate the flow patterns in the dam, the team obtained drawings from LCRA 19

23 that provided a better understanding of the geometry of the tunnels in the dam. LCRA blueprint 249-D-1352 indicates that the original blower in the first dehumidification system was 4.2 m 3 /s (9000 cfm). The design team had opportunities to closely examine the air flow paths in the tunnels and connecting ventilation passages on July 11, The design team walked through the gate galleries, inspection galleries, and drainage galleries observing air flow at tunnel junctions, inlet air vents, and ventilation outlets. Based on this examination, the team was able to estimate the pattern in which air flowed through the dam. Although this cursory analysis offered insight into the basic air flow pattern in the dam, the team decided that more analysis, such as a numerical model of the air flow pattern in the dam, was necessary to specify the performance of a dehumidification system. The design team directed the majority of this analysis at the gate galleries rather than the entire network of tunnels in the dam. LCRA concurred with the decision to focus on dehumidifying the gate galleries because most of the equipment requiring maintenance due to corrosion is located in these tunnels. The team could have designed a system to dehumidify every tunnel in the dam, but the additional cost to install and operate such a system would yield little benefit in reduced maintenance costs. After deciding to focus on the gate operating galleries, the team began a numerical analysis of the air flow patterns in the dam. The team treated the tunnels as a network of pipes and analyzed the flow using equations developed by professor Hardy Cross [Daugherty, 1977]. Calculations for pressure drops through the galleries were taken from the American Society of Heating, Refrigerating and Air Conditioning Engineers 20

24 (ASHRAE) Handbook, Fundamentals [ASHRAE, 1989]. A detailed explanation of this analysis is presented in Appendix C. Figure 10 presents a schematic of the tunnels containing equipment used to raise and lower the floodgates in Mansfield Dam. Each tunnel is assigned a number for ease of reference. Each flood gate consists of two gates: an upstream gate and a downstream gate. These gates are represented as shaded boxes in Figure 10. Arrows indicate the direction of air flow in the tunnels. Dehumidification Bunker 8 Downstream Gates 1-6 Downstream Gates Upstream Gates Upstream Gates 7-24 Figure 10: Gate operating galleries. Figure 11 presents the results of the design team s air flow model. This particular simulation is based on a system that processes nine thousand cubic feet of air per minute. Dehumidification Bunker 9000 cfm Downstream Gates 1-6 Downstream Gates cfm 4600 cfm 5200 cfm 4400 cfm 600 cfm 3800 cfm 4400 cfm Upstream cfm Upstream

25 In examining the air flow model results presented in Figure 11, one can see that the downstream galleries receive more air than the upstream galleries. The air flow rates calculated using the design team s model are used in the evaporation model presented in the next section. These flow rates are also used in the calculations to determine how long it will take the dehumidification system to dry out the dam Evaporation From Trenches and Walls The design team recognized the importance of quantifying the amount of water that enters the air inside the dam by means of evaporation. This quantity dictates the mass flow rate and humidity ratio of the air a dehumidification system must supply. Wet surfaces exist throughout each tunnel inside the dam. Some of these surfaces, such as the floodgate machinery, are wet as a result of condensation and will no longer act as a source of evaporating water once a dehumidification system has reduced humidity levels in the dam. Other surfaces, such as water-filled drainage trenches that run along each tunnel and walls wetted by water seepage through expansion joints, will remain wet. Water will continue to evaporate from these surfaces even after a dehumidification system is functioning. The rate of evaporation at any point along a tunnel strongly depends on humidity, 22

26 flow rate, and wetted surface area at that specific location. Lower air humidity, higher flow rates, and greater wetted surface area all contribute to a higher evaporation rate. Humidity is the easiest of the three parameters to determine. After designing a system, the design team knew what the humidity of the air leaving the system and entering the tunnels would be. The team then iterated to find the humidity in the tunnels based on the amount of wetted surface area over which the air flowed. Flow rate was more difficult to determine. The design team knew the flow rate exiting the dehumidification system, but, as mentioned in Section 3.1.2, the team had to understand the flow patterns through the gate galleries in order to know the velocity at every point inside the tunnels. These air velocities were then used in the evaporation model to generate the maximum allowable absolute humidity at the entrance to the gate gallery tunnels. This value was used to specify the absolute humidity at the output of the dehumidification system. The wetted surface area was the last of the three parameters that had to be quantified. The design team used the width of the water-filled trenches as an estimate of the wetted surface area per unit length for each tunnel. This estimate is a compromise between the current wetted surface area and the steady state wetted surface area. With no dehumidification system in operation, the present wetted surface area is much more extensive than the surfaces of the water in the trenches. After a dehumidification system is functioning, however, these wetted surfaces will evaporate, leaving only the water in the trenches. All of the trenches in a given level of the dam are inter-connected, which means that water may flow into the gate gallery trenches from adjacent tunnels. If this water flows into the gate gallery trenches at a rate high enough to make up for the evaporation, then the trenches will remain wet. If the evaporation rate matches or 23

27 exceeds the inflow rate, then portions of the gate gallery tunnels will become dry. In this case, the total wetted surface area in the gate gallery tunnels will be less than the value used in the evaporation rate calculations, and the humidity specifications based on this model represent an overdesign. The design team based the estimate of the wetted surface area on the assumption that the tunnels will remain wet so that any error in this assumption is in the direction of overdesign. The team prefers the risk of overdesigning to the risk of specifying a system that cannot meet the dehumidification requirements. The design team developed an evaporation model that incorporates the above parameters to calculate the changing humidity ratio through each tunnel. This model is based on the convection mass transfer relations for internal flow presented in Fundamentals of Heat and Mass Transfer [Incropera, 1996]. This model incorporates the design team s best guesses of initial humidity ratio, air velocity, and wetted surface area to perform the calculations. The accuracy of the model improves with better estimates of these three parameters. This model, the governing mass transfer relations, and a summary of the output are described in detail in Appendix D. The model indicates that treated air flowing through the tunnels asymptotically approaches a saturated condition. This behavior is a result of a decreasing evaporation rate that corresponds to an increasing humidity ratio. As air passing through a tunnel absorbs water, the species concentration of the water vapor in the air increases. This increase yields a drop in the species concentration differential between the surface water and the moving air. This concentration differential is the driving force behind evaporation. As the species concentration differential decreases, the rate of evaporation decreases as well. Because the humidity is highest at the points farthest from the entrance, these points represent the worst case locations which will dictate the 24

28 specifications of any proposed dehumidification solution. Once operational, any dehumidification system must maintain an acceptable humidity ratio within the dam. This specified humidity ratio will ensure that water does not condense on cool surfaces and that all of the tunnel surfaces will remain dry. Upon initial startup, however, the dehumidification system must be capable of removing the water that is currently inside the dam within a reasonable time period. The system must also dry the tunnels after any periods of down time if the system ever fails. To calculate the transient drying time of the dehumidification system, the design team calculated the rate at which water will be removed by the moving air and estimated the amount of water currently inside the gate gallery tunnel circuit illustrated in Figure 10. The following equations outline the method employed by the design team to arrive at an estimate for the transient drying time. water on tunnel surfaces = ( tunnel lengths tunnel perimeters) wetted percentage water layer thickness ( ω ω ) rate water is removed = m out in air drying time = water on tunnel surfaces rate water is removed The air flow circuit tunnel lengths multiplied by their respective perimeters yields a surface area of approximately 6500 m 2. The team used 50 % as a conservative estimate of the wetted percentage, which makes the wetted surface area approximately 3250 m 2. The design team then assumed that the water layer on the wetted surfaces is roughly 1 25

29 mm thick. This makes the estimate for the total water volume on the tunnel surfaces equal to 3.25 m 3, which corresponds to a mass of 3250 kg. The rate at which water is removed from the tunnels by the air is directly related to the difference between the incoming and outgoing humidity ratios. The input humidity ratio to the tunnel circuit is the same as the output humidity ratio specified for the dehumidification system, which is kg water /kg air. During this drying period, the humidity ratio at the exit of the tunnel circuit will be the humidity ratio that corresponds to saturation at the exit temperature. This humidity ratio is approximately kg water /kg air. The difference in humidity ratios is then kg water /kg air. The specified air flow rate is 9000 ft 3 /min, or 4.25 m 3 /s. This flow rate corresponds to a mass flow rate of 5.1 kg/s which is multiplied by the difference in humidity ratios to find the rate at which water can be removed by the system. This rate is 0.01 kg water /s. The time taken to dry the tunnels is the amount of water in the tunnels divided by the rate at which the water is removed. This calculation yields a time of approximately 90 hours, or roughly 4 days. The design team feels that this is a very appropriate drying time. A shorter time would indicate that the system is over designed, and any system requiring a drying time significantly longer than 4 days is probably inadequate. This estimate of the transient drying time is based on a number of rough estimates, including the total tunnel surface area, the average thickness of the water on those surfaces, and the percentage of those surfaces that are wet. The above transient calculation is informative in that it provides an order-of-magnitude estimate of the drying time. The calculation shows that the gate galleries will dry out in several days, as 26

30 opposed to several hours or several weeks. This estimate is accurate enough to indicate that a dehumidification system that meets the given specification will dry the tunnels in a reasonable period of time. The reader is expected, however, to understand that these are order-of-magnitude calculations Necessary Conditions of Treated Air Once the team modeled the mechanisms that remove and introduce moisture to the air in the dam tunnels, the team determined the required conditions of air exiting the dehumidification system. The conditioned air has to have a dew point lower than the temperature of the coldest surfaces in the tunnels. The air must be able to absorb a given amount of moisture from surfaces in the tunnels and still resist condensing on cool machinery at points farthest from the conditioning system. The team identified evaporation in the tunnels, air flow patterns through the dam, and temperature of floodgate machinery in the dam as parameters that govern moisture levels in the dam. The amount of wetted surface area inside the dam also has a significant effect. Once these parameters were quantified, the team calculated the required temperature, humidity, and flow rate of air exiting the dehumidification system. Calculations of the time to dry out the dam indicate a required flow rate of 4.25 m 3 /s (9000 cfm). The air flow and evaporation models presented in Appendices C and D require that the temperature of the air entering the dam be 21 C (70 F) with a dew point of 10 C (50 F) Conditions of Outside Air 27

31 The dehumidification system has to condition the warm, moist air drawn from the surroundings. Temperature and humidity data for ambient air was used to determine the amount of water in the air entering the dehumidification system. Given an initial amount of water in the air and necessary conditions for air exiting the dehumidification system, the design team was able to determine how much moisture the dehumidification system has to remove from the air. The team obtained 1% design temperatures, or highest wet and dry bulb temperatures that statistically may occur in one out of 100 days, for Austin, Texas. These design temperatures signify the worst case conditions in which the new dehumidification equipment must operate. The given dry bulb temperature of 37.8 C (100 F) and wet bulb temperature of 25.6 C (78 F) yield an absolute humidity of kg water /kg dry air [ASHRAE, 1989] Water Temperature at System Entrance The temperature of the water on the upstream side of Mansfield Dam is the temperature of a possible cooling medium in a dehumidification system. The existing system used water bled from a floodgate to cool the air. During the summer months, the temperatures at the bottom of the dam are lower than the temperatures closer to the water's surface. As was mentioned above in section 3.1.1, on June 23, 1997 the team measured the water coming out of a floodgate and found it to be 18.3 C (64.9 F). The tail race temperatures in Appendix A and temperature records from a limnology reference were useful for comparing possible floodgate temperatures and water entering the 28

32 dehumidification system. The team examined this temperature data in order to identify seasonal variations in the lake water temperatures. This data was used both to confirm floodgate temperatures and to generate possible specifications for any water cooled dehumidification equipment that the team considered. The most conservative temperatures from the perspective of the dehumidification system were the warmer temperatures. The warmest temperature trends appeared in the tail race data. However, this data varies greatly due to droughts and flooding in recent years. Other data taken from a limnology report of 1963 through 1965 was used as typical year data to specify the dehumidification system. The tail race temperature information served to test system feasibility in unusually dry and wet weather. Based on this analysis, the design team used 18 C (64 F) for the temperature of cooling water fed to the dehumidification system Performance of Conditioning System Once the parameters in Sections through were quantified, the design team incorporated them into a specification sheet, shown in Table 2. The specification sheet details the performance required of the final design. This information was useful to the design team for the consistent development of different concept variants. One of the parameters on the specification sheet is absolute humidity. Success of the concept variants was measured by this dimensionless ratio. The systems were considered feasible if they could produce air with an absolute humidity of to kg water /kg dry air at the exit of the dam. This means that condensation would never appear 29

33 on the inner surfaces of the dam. In addition to eliminating condensation by achieving a low absolute humidity, the feasible dehumidification systems had to conform to other constraints. 30

34 Table 2. Specification Sheet for Dehumidification System, Mansfield Dam. Design Specifications Value of Specification Volume of Treated Space 2265 m 3 (80000 ft 3 ) Design Dry Bulb Temperature (Input) Design Wet Bulb Temperature (Input) Minimum Design Dry Bulb Temperature (Output) Maximum Design Dry Bulb Temperature (Output) Design Humidity Ratio (Inlet) Design Humidity Ratio (Equipment Outlet) Design Humidity Ratio (Treated Space Exit) Treated Space Design Dew Point 37.8 C (100 F) 25.6 C (78 F) 19 C (66 F) 27 C (81 F) kg water /kg dry air (113 grains water /lb dry air ) kg water /kg dry air (63 grains water /lb dry air ) kg water /kg dry air (77 grains water /lb dry air ) 15.6 C (60 F) Number of Exchanges per Hour 6.75 Volumetric Flow Rate Maximum Design Temperature of Available Cooling Fluid Available Electricity Largest Allowed Cross-Section of Components (Corresponds to Bunker Door) Dimensions of Available Space with Modifications in Bunker (L x W x H) Dimensions of Available Space without Modifications in Bunker (L x W x H) Area of Metal Access Grate in Bunker (Largest Allowable Component Size if no Construction) 4.25 m 3 /s (9000 cfm) 10 C (50 F) 480 V, 3 phase and 120 V, single phase 1.5 m x 2.3 m (5 ft x 7.5 ft) 4.26 m x 2.74 m x 3.04 m (168 in x 108 in x 120 in) 1.8 m x 1.7 m x 3.0 m and 3.3 m x 2.7 m x 2.1 m (72 in x 66 in x 120 in and 132 in x 108 in x 84 in) 1.2 m x 1.8 m (47 in x 71 in) 31

35 Note that the specification sheet is for the dehumidification system, not just the dehumidification equipment. The specification sheet includes system specifications that are not unique to different concept variants. For example, the volume of the space to be conditioned within the dam is a specification that describes any proposed system. The available electricity is specified, 480 V and 120 V. The dimensions of the available space for dehumidification equipment are also specified, both with and without modifications made to the existing dehumidification bunker. 3.2 Functional Analysis Figure 12 shows the functional analysis that the design team conducted. This analysis revealed several functions that the team had to consider in the dehumidification system design. The solid line boxes contain system functions. The solid boxes are functions that all systems must have in common. For example, any possible dehumidification system has electricity available in the dehumidification bunker, so the system has to be equipped to receive electricity. Also, the system must be able to filter air, move air, and dehumidify air. The dashed boxes contain functions that are optional, like pre-cooling and post-cooling. Boxes with gray bases contain two independent, parallel functions: fit in bunker and secure equipment. These functions indicate that the system is physically limited by the geometry of the existing bunker and has installation constraints. The double box outside of the system boundary signifies a user function. The team made the assumption that an LCRA employee would regularly clean the system filters. 32

36

37 The arrows signify energy, material, and signal flow between functions. The thin Figure 12. Function structure. solid arrows represent energy flows like electrical, thermal, and human energy. The thick solid arrows represent material flows, generally water and air at various states. The dashed arrows are signal flows between system functions which indicate where there is a need for system controls. The values T1, T2, T3, and T4 indicate temperatures of flows of air and water. 3.3 Solution Principles The team identified several solution principles to consider while designing the dehumidification system. These solution principles cover systems that utilize the cool lake water, such as the existing system, to systems based on a vapor-compression (VC) cycle. The team also considered several hybrid systems that utilize both the cool lake water as well as a vapor-compression cycle. Table 3 presents these solution principles. The Comments column of Table 3 identifies some potential advantages and disadvantages associated with each of the concept variants. The design team had several solutions to compare. Then the systems' feasibilities had to be checked against the performance criteria set by the team. The recommended system will remove all condensation from the inner surfaces of the dam, particularly the floodgate equipment. With inner gate gallery surfaces at 15.5 C (60 F) or warmer, surfaces will be kept dry if the dehumidification equipment maintains a dewpoint of 10 C (50 F). 34

38 Table 3. Potential Concept Variants. System Type Operation Principles Comments 1 Current system Uses cool lake water to cool and dehumidify air entering dam. Lake water is a free and abundant resource. Some of the existing equipment may be salvageable. The lake water may not be cool enough for the system to work 2 Vaporcompression system 3 Water pre-cooled vaporcompression system 4 Vaporcompression with water cooled condenser 5 Water pre-cooled vaporcompression system with water cooled condenser Cools and dehumidifies air by rejecting heat to the environment. Combination of systems 1 and 2. Similar to system 2, but rejects heat to the cool lake water instead of to the warm environment. Combination of systems 3 and 4. 6 Desiccant system Uses moistureabsorbing properties of desiccants to dehumidify air. properly. Such systems are common and fairly inexpensive. The compressor requires considerable energy input. Uses lake water to reduce cooling load of VC system, reducing required energy input. High initial cost. Rejecting heat to the cool lake water reduces the required pressure differential within the VC system, which results in less required input energy. High initial cost. This system utilizes the cool lake water for two purposes, resulting in the lowest energy requirement of all the VC systems. Very high initial cost. Requires very little energy input. Requires frequent maintenance. No estimate yet on cost System Using 100% Outside Air Based upon the initial understanding of the dehumidification problem, the design team chose to focus on systems that use 100% outside air. This type of system pulls ambient air into the dam, treats the air, and directs this air through the dam. The treated air passes once over inner surfaces and then exits the dam. The original dehumidification 35

39 equipment operated in this manner. The team considered rehabilitation of the old system, installation of a vapor-compression system, installation of a desiccant system, and combinations of these system. All of these concept variants operate by using 100% outside air. Using outside air in a single pass through the dam has the added advantage of cooling the power house. The power house has two blowers that pull air from a passage in the dam downstream of the gate galleries. The air that has passed through the gate galleries is cooler that the air in the power house, so using this resource would be very economical for LCRA. Another advantage of the 100% outside air configuration is that no ductwork needs to be installed within the dam, and no construction that would interfere with present normal dam operations is necessary. The tunnels themselves form the ductwork. Also, the original dehumidification bunker can continue to house the equipment. A disadvantage is that dehumidification systems using 100% outside air have to be sized large enough to continuously dehumidify ambient air. Some modifications may be required to allow new equipment to fit inside the bunker Closed Loop System The design team also considered the possibility of installing a system that does not use 100% outside air. This system involves using strategically placed doors to close off the circuit of tunnels that houses floodgate equipment. A small dehumidifier and fan condition and recirculate dehumidified air in the circuit. This idea was proposed late in 36

40 the project analysis, so the design team focused primarily on the concept variants that did not involve closing off any tunnels. The design team is calling this arrangement a closed loop system because doors would be used to seal the gate galleries off from the outside air thus forcing dehumidified air to cycle continuously through the gate galleries. An example of this cycle is shown in Figure 13. Sealed doorways Dehumidification Bunker Figure 13. Closed loop example. Four doorways that can be sealed control the inlet and outlet passages 1 and 8. Tunnel 5 is also sealed off by doors. Dehumidification equipment, which includes a blower or fan, directs the flow of treated air through tunnels 2, 3, 6, 9, 7, and 4 in a continuous cycle. This system would work like a residential air conditioning unit in that the same treated air is cycled through the system, and small amounts of outside air can be brought into the system for ventilation. 37

41 Advantages of this arrangement include having smaller dehumidification equipment with smaller volumetric flow requirements than dehumidification equipment treating 100% outside air. The treated air from the closed loop system stays in the tunnels, so the equipment can continue to treat the air after each pass through the gate galleries. Moisture would continue to be drawn out of the tunnels with the same air. However, the power house cannot then be cooled using this system. Other problems with the closed loop configuration include the lack of usable space within the gate galleries for dehumidification equipment. The tunnels have to remain clear to allow the movement of hoist and tackle equipment. Occasionally, the flood gate equipment is taken out of position and moved outside through the tunnels for maintenance. The doorways would also have to be removed and reinstalled before and after each time flood gate equipment moves in or out of the gate galleries. IV. Concept Variants The design team investigated the feasibility of four concept variants. 4.1 Water Cooled Heat Exchanger The design team s first concept variant consists of rehabilitating the original dehumidification equipment within Mansfield Dam. Figure 14 presents a schematic of this system. 38

42 warm moist air cool lake water filter heat exchanger heat added if necessary blower cool, drier air warmed lake water condensate water discharge Figure 14. Water cooled heat exchanger concept variant Principles of Operation The principles of operation of the water cooled heat exchanger system are expressed in more detail in Section 2.3. This dehumidification system reduced the moisture content in the air by passing the air through a heat exchanger that was cooled by lake water. Heaters were also installed in the original system to reheat the air to comfortable temperatures before the air was directed into the dam Feasibility The components in the existing dehumidification system are not currently functional. Some of the filters are missing. One of the heads and several of the tubes in the heat exchanger are severely damaged. The blower is out of balance and may not function at all. Many of the resistive heating elements in the reheater are missing. In addition to these missing and damaged components, both the electrical system and the plumbing system would have to be re-evaluated and possibly replaced before the system 39

43 could function. The design team has identified vendors that can provide replacements for the major components in the existing system, but the team eliminated this concept variant from further consideration based on feasibility analysis. The water cooled heat exchanger can only cool the air to the temperature of the lake water entering the system. This water is at a temperature of approximately 18 C (64 F), so the heat exchanger cannot cool the air to the specified dewpoint temperature of 10 C (50 F). Therefore, this system fails the criterion of lowering the humidity to an acceptable level. This concept variant is still attractive, however, because the cooling medium, lake water, can be utilized at no cost. Although this system cannot prevent condensation on machinery in the dam, the lake water cooled heat exchanger is a free low temperature heat sink that the design team can incorporate into other designs. 4.2 Air Cooled Vapor-Compression The second concept variant that the design team considered is an air cooled vaporcompression system. This type of dehumidification system utilizes a standard vaporcompression cycle to remove humidity from the air Principles of Operation The second concept that the design team considered was an air cooled vaporcompression system. This system removes humidity by cooling the air, which is the same principle on which the water cooled heat exchanger operates. Unlike the water cooled 40

44 heat exchanger, however, a vapor-compression system takes advantage of the phase change properties of a working fluid to absorb heat from the air. This technique allows vapor-compression systems to cool air to lower temperatures than the water cooled system. Vapor-compression systems are capable of cooling the air to temperatures just above 0 C. This limitation is imposed by the tendency of water to freeze on the heat exchanger coils at such temperatures. The vapor-compression system is outlined in Figure 15. The system contains a compressor, a condenser, a throttling valve, and an evaporator. The working fluid enters the compressor in a gaseous phase and at a low pressure. The compressor increases the pressure of the fluid and forces it into the condenser. At this high pressure, the working fluid will condense even though the condenser is relatively warm. The heat generated by this phase change is rejected to the warm ambient air. The fluid is in a liquid state as it leaves the condenser and moves to the throttling valve. The throttling valve serves to maintain the pressure differential generated by the compressor. The working fluid exits the throttling valve at a low pressure but still in the liquid state. At this low pressure, the liquid will evaporate at relatively low temperatures. The evaporator transfers heat from the moving air to the working fluid to drive this evaporation process. This heat transfer from the moving air is the means by which the air is cooled to the necessary temperature, 10 C (50 F) [Howell, 1992]. 41

45 warm moist air condenser hot air rejected throttling valve compressor (work input) evaporator warm moist air cool dry air Figure 15. Air cooled vapor-compression concept variant. The air cooled VC system is already in industrial use to dehumidify environments similar to the interior of Mansfield Dam such as basements, pumping stations, and wastewater plants. Dew points of 4 C (40 F) can be achieved, more than enough to sufficiently reduce humidity in the dam [Dehumidification, 1996] Feasibility To meet the specifications outlined by the design team, the dehumidification system must change the state of the air from the specified ambient conditions to the specified outlet conditions. This change of state requires a considerable amount of energy. This energy requirement is calculated below. where: Energy = m air h air 42

46 m = 51. h kg s kj = 50 kg Energy = 255 kw The above calculation indicates that a cooling-based dehumidification system must have a capacity to transfer heat from the moving air to the environment at a rate of 255 kw. This is equivalent to a cooling capacity of 72 tons. This cooling requirement can be lowered significantly with the introduction of a water cooled heat exchanger, as described in Section 4.1. The water cooled heat exchanger can pre-cool the incoming air to approximately 18 C (65 F). This changes the h requirement to 22 kj/kg, which reduces the required capacity to 112 kw, or 32 tons. This configuration is shown in Figure 16. hot, humid air (from environment) condenser hot air rejected to environment cool water from lake throttling valve compressor (work input) heat exchanger hot, humid air (from environment) warm moist air evaporator cold dry air heater cool dry air routed to dam warmer lake water rejected to environment Figure 16. Air cooled vapor-compression concept variant with precooling. 43

47 4.2.3 Available Systems The design team was unable to consolidate vendor information for any of the vapor-compression systems because the contacted vendors did not respond to the team s requests for information in a timely manner. The team was able, however, to make some fair approximations based on some data that was received and on data obtained through other sources. One dealer of vapor-compression systems provided an estimate for a 40 ton capacity unit for $12,300 [Rodak, 7/16/97]. This estimate was for a vapor-compression system without any pre-cooling capability. The calculation in the previous section indicates that the required tonnage for a system with pre-cooling is approximately 32 tons. If we use the estimate as a scaling factor, the cost of a 32 ton unit should be approximately $9,900. An estimate for a water cooled heat exchanger that can serve as a pre-cooler was provided by a different vendor as $22,500. The design team feels that this estimate is high, but it is the only data point available, and the team must base the estimations on the data at hand. The cold dry air exiting the vapor-compression system must be reheated to a comfortable temperature before being introduced to the dam interior. The air must be heated from 10 C (50 F) to 21 C (70 F). For the mass flow rate specified by the design team, this reheat requirement will be 57 kw. The team obtained a price quotation for a 60 kw capacity reheat system costing $7,700. The total estimated cost of the vaporcompression system, the pre-cooler, and the reheat system is $40,100. The operating cost associated with the above system consists of electricity costs for the vapor-compression system and for the reheat system. Typical air cooled vaporcompression systems have a coefficient of performance (COP) of 3.2 [Trane, 7/30/97]. The COP is defined as the heat transfer rate divided by the work input rate [Howell, 44

48 1992]. The above system has a capacity of 112 kw (32 tons), so the input power requirement should be approximately 35 kw. This cost is added to the reheat power consumption of 57 kw to arrive at a total operating power consumption rate of 92 kw. This value and the estimated cost value are used in the economic analysis in Appendix E. 4.3 Water Cooled Vapor-Compression System The third concept variant, the water cooled vapor-compression system, also incorporates a vapor-compression system Principles of Operation Unlike the air cooled vapor-compression system, this system, presented in Figure 17, uses cool lake water to cool the condenser instead of warm ambient air. Because the water cooled condenser is cooler than the air cooled condenser, the compressor does not have to increase the pressure of the working fluid as much to make it condense. Therefore, the water cooled system requires less energy to operate and is more efficient than the air cooled system. cool lake water condenser warmer lake water rejected throttling valve compressor (work input) evaporator warm moist air cool dry air Figure 17. Water cooled vapor-compression concept variant. 45

49 4.3.2 Feasibility The water cooled vapor-compression system is identical to the air cooled vapor compression system except that the condenser is cooled with water instead of air. Like the air cooled system, the required capacity can be greatly reduced by incorporating a water cooled pre-cooler into the system. Like the air cooled system, the introduction of the pre-cooler reduces the required capacity to 112 kw (32 tons). This configuration is shown in Figure 18. cool water from lake condenser warmer lake water rejected to environment throttling valve compressor (work input) hot, humid air (from environment ) heat exchanger warm moist air warmer lake water rejected to environment evaporator heater cool dry air routed to dam Figure 18. Water cooled vapor-compression concept variant with precooling Available Systems The water cooled vapor-compression system is similar enough to the air cooled system that the design team based the cost estimate on the same data. The actual 46

50 condenser in a water cooled system may cost a little more than an air cooled condenser, because the water cooled technology is somewhat less common. The compressor in a water cooled system is required to do less work, though, so the reduced cost of the compressor may counteract the possible increased cost of the condenser to bring the total costs of the two systems to the same level. Therefore, the design team assumes that the equipment cost for the water cooled vapor-compression system is approximately $40,100. The operating cost of the water cooled vapor-compression system is less than that of the air cooled system. Typical COP values for water cooled systems are 4.2 [Trane, 7/30/97]. To meet the heat transfer requirement of 112 kw (32 tons), the water cooled vapor-compression system should require approximately 27 kw. The total power consumption of the water cooled vapor-compression system plus the reheat system is 83 kw. This value is used in the economic analysis presented in Appendix E. 4.4 Desiccant Wheel The design team s fourth concept variant uses a desiccant wheel to dehumidify the air. Figure 19 presents a schematic of the desiccant wheel system. hot dry air (to dam) rotation warm moist air desiccant wheel partition hot air to recharge desiccant (energy input) moist hot air rejected Figure 19. Desiccant wheel concept variant. 47

51 4.4.1 Principles of Operation The desiccant wheel dehumidification system uses desiccants to draw moisture out of the air. A desiccant system is capable of producing the driest air of team s concept variants. There are many kinds of desiccant materials and system configurations. Desiccants can be in liquid or solid form. They can be regenerated by heat or disposed of after each use. Desiccants can also be absorbent or adsorbent materials. Absorbents react with water chemically, and adsorbents react physically [James, 1973]. The design team is considering a type of desiccant dehumidification system that is commonly used for storage and packaging applications [Dryomatic, 1965], the silica gel desiccant wheel dehumidification system. Silica gel is an adsorbent material, which means that water molecules attach to the desiccant without causing a chemical change in the material. This type of desiccant, silica gel, can also be regenerated by heating to cause the desiccant to give up water. The temperature required for regenerative heating varies, but complete regeneration without pre-cooling or post-cooling may require temperatures of up to C ( F) [Dryomatic, 1965]. Wheel desiccant systems are capable of achieving very low dew points, down to -40 C [Dehumidification, 1996]. Vapor-compression systems rarely reach dew points this low because water begins to freeze on the condenser coils at 0 C (32 F). However, desiccant systems have a high initial cost and would expel very hot, dry air into the dam tunnels. The main component of this system is a wheel containing silica gel that is approximately 1.5 m (5 ft) in diameter and rotating at about 3 to 4 revolutions per hour [CT, R & RC Series]. The wheel is partitioned to allow two separate air flows to pass 48

52 through it. One flow of moist air drawn from the environment passes through half of the wheel. The desiccants in the wheel remove moisture from this air, and the flow is then directed into the dam. A second flow of hot, dry air is used to recharge the silica gel in the wheel. As this hot, dry air passes through the other half of the wheel, it removes moisture from the silica gel. The wheel rotates to allow for periodic recharging of the silica gel. Any given point on the wheel spends a certain amount of time in both air flows. After the moist air has saturated the silica, it rotates to the opposite side and is recharged by the hot, dry air [Ingram, 1982] Feasibility The design team has determined that a desiccant wheel dehumidification system is a feasible solution for the humidity problems in Mansfield Dam. A system with a flow rate of 4.25 m 3 /s (9000 cfm) is necessary to have a short response time to dry out the dam (see Section 3.1.3). Desiccant wheel dehumidification systems can have volumetric flow rates of m 3 /s to more than 18.9 m 3 /s (10 cfm to more than 40,000 cfm) [Munters, 1990]. Wheel desiccants with a volumetric flow of 4.25 m 3 /s have a nominal face velocity of 3.05 m/s (600 fpm). Face velocity is the speed of reactivation air through the wheel. With this velocity the desiccant wheel can treat air from kg water /kg dry air to kg water /kg dry air (115 grains water /lb dry air to 58 grains water /lb dry air ) [ATS, 1995]. This is more than enough capacity to remove moisture from kg water /kg dry air to kg water /kg dry air (113 grains water /lb dry air to 63 grains water /lb dry air ). In addition to being able to meet the requirement of removing moisture from air, the desiccant wheel system must fit in the available space in the dehumidification bunker. The size of desiccant systems is driven by desiccant wheel diameter and blower capacity. Systems with volumetric capacities of 4.25 m 3 /s (9000 cfm) have wheels approaching

53 m (5.5 ft) in diameter [LaRoche, 1997]. The wheel by itself approaches the limits of the space available in the bunker, but this concept variant is feasible if modifications are made in the bunker to remove old equipment. These modifications result in an available space of 4.26 m x 2.74 m x 3.04 m (168 in x 108 in x 120 in) into which a desiccant wheel dehumidifier with a volumetric flow rate of 4.25 m 3 /s (9000 cfm) could fit Available Systems The design team investigated four different desiccant wheel systems. Two systems were found to be feasible. The feasible systems are the LHS-9000 from Low Humidity Systems (LHS) and the HCD-9000 EA from Cargocaire. The other two systems failed to meet the team s size requirements. The DES E/S from Desicair is too long at 4.7 m (185 in) in length and the AFD from Airflow is also too long at 8.5 m (336 in) in length. Table 4 describes the two feasible systems. Table 4. Two embodiments, desiccant system LHS-9000 HCD-9000 EA Minimum Flow Rate 3.54 m 3 /s (7500 cfm) 2.12 m 3 /s (4500 cfm) Maximum Flow Rate 5.31 m 3 /s (11250 cfm) 5.66 m 3 /s (12000 cfm) Length 3.5 m (137 in) 3.3 m (129 in) Width 2.4 m (94 in) 2.01 m (79 in) Height 2.4 m (95 in) 2.26 m (89 in) Weight 885 kg (1950 lb) 4055 kg (8940 lb) Lowest Dew Point ** -48 C (-55 F) Electrical Requirements 480 V 480 V Power Consumption ** 205 kw Unit Cost $38,000 $60,000 Heat Exchanger Cost $22,500 not necessary Combined Cost $60,500 $60,000 ** Values are assumed to be equal to those for HCD-9000 EA. The LHS-9000 has an adjustable volumetric flow rate from 3.54 m 3 /s to 5.31 m 3 /s 50

54 (7500 cfm to cfm), has dimensions (length by width by height) of 3.5 m x 2.4 m x 2.4 m (137 in x 94 in x 95 in) and weighs 885 kg (1950 lb) [Adams, 7/15/97]. The HCD EA has an adjustable volumetric flow rate from 2.12 m 3 /s to 5.66 m 3 /s (4500 cfm to cfm), has dimensions (length by width by height) of 3.3 m x 2.01 m x 2.26 m (129 in x 79 in x 89 in), and weighs 4055 kg (8940 lb) [Munters, 1996]. Both of these systems can run on 480 V, which is available in Mansfield Dam [Sumner, 7/17/97]. The LHS without post-cooling costs approximately $38,000 [Denny, 7/30.97]. A heat exchanger for post-cooling can be purchased from O & M Manufacturing for about $22,500 [Cash, 7/30/97]. The total price of the LHS-9000 system is then $60,500. The HCD-9000 EA with post-cooling included costs approximately $60,000. The design team does not know the power consumption for the LHS-9000 system but does know that the power consumption for the HCD-9000 EA is 205 kw for continuous operation [Moreland, 7/28/97]. The team is assuming that the power consumption for both systems is approximately equal. For more specific vendor information, see Appendix F. V. Analysis of Concept Variants Four concept variants passed through the design team s feasibility analysis. These are the air cooled vapor-compression system, the water cooled vapor-compression system, the LHS-9000 desiccant wheel system, and the HCD-9000 EA desiccant wheel system. The design team made a final recommendation by using two tools called an objective tree and a decision matrix. An objective tree is a graphical schematic that 51

55 assigns an importance value to system parameters. All of the important values are percentages that total to 100%. The decision matrix arranges the concept variants and system parameters in a matrix configuration. Each concept variant is then scored on its performance in relation to the system parameters. The values from the objective tree are used to scale each concept variant s score. The highest score is selected as the final recommendation. 5.1 Objective Tree The design team developed a formal plan for evaluating the various concept variants based on performance and cost. The team created an objective tree to determine the relative impact of different parameters on the final decision. Figure 20 shows this objective tree, which indicates the relative importance the design team assigned to the various parameters. System Whole Value: 100% Present Worth Value: 50% Performance Value: 50% Humidity Transient Capacity Comfort Value: 15% Value: 20% Value: 15% Figure 20. Objective tree for determining relative importance of system parameters. 52

56 The present worth is a combination of initial costs, long-term costs, and long-term benefits. Initial costs include the purchase price as well as all other costs associated with installing the system and bringing it into operation. Long-term costs include operation and maintenance costs. The long-term benefit of a given system is an estimate of the present value of the expected savings that will be generated by the system over its lifetime. These savings will result from reduced maintenance on the floodgate equipment and also represent savings related to improved safety and employee comfort. This longterm benefit is directly related to system life because the savings are realized each year the system functions. Therefore, the present worth of each system is primarily a function of expected system life. The design team used a life of 20 years for present worth calculations. Conversations with engineers in the Heating, Ventilation, and Air Conditioning (HVAC) industry confirm this value as a good estimate [Moreland, 7/25/97 and Bodine, 7/25/97]. Appendix E presents a complete economic analysis of all of the concept variants. This appendix shows that, for operating lives of 10 to 20 years, vaporcompression systems offer three to four times the savings of desiccant systems. The concept variant with the highest present worth is awarded the full fifty points in that category. Any concept variant with a present worth of zero at 20 years is given zero points. The other concept variants are assigned points consistent with this scale. Humidity is the first performance parameter upon which the concept variants are ranked. All systems must meet the specifications outlined in Section 3.1.7, but extra credit is assigned to any system that can provide a humidity ratio below the specified 53

57 value of kg water /kg air (63 grains water /lb dry air ). A system providing a humidity ratio of kg water /kg air is given zero points in this category, and two extra points are assigned for each kg water /kg air (7 grains water /lb dry air ) below that value. Therefore, a system with an output humidity ratio of kg water /kg air (11 grains water /lb dry air ) would receive the full 15 points in this category. The second performance parameter listed is Transient Capacity. As discussed in Section 3.1.3, any dehumidification system will require some time to initially dry the tunnels in the gate gallery level within the dam. A system that meets the design team s specifications should dry these tunnels in approximately four days. Any system that can dry the tunnels in less time is given extra points. The transient response time is based on both air flow rate and humidity ratio. Because humidity ratio is the basis for points in the previous category, that parameter is filtered from this category. Therefore, points are awarded in this category solely on the basis of air flow capacity. A system meeting the specified flow rate of 4.25 m 3 /s (9000 cfm) is given no points. One point is then assigned to a system for every 0.24 m 3 /s (500 cfm) in capacity beyond 4.25 m 3 /s (9000 cfm). A dehumidification system providing air at m 3 /s (19000 cfm) would be given the full 20 points in this category. Comfort is the last performance parameter upon which the final decision is based. LCRA employees must enter the dam regularly for maintenance, and the design team believes that the comfort of those employees warrants some consideration in the overall decision. A temperature of 23 C (73 F) is specified as the ideal temperature for human comfort [ASHRAE, 1989]. A dehumidification system that outputs air at that 54

58 temperature is assigned the full 15 points. A system loses 3 points for each degree ( C) of difference between 23 C and the actual temperature of its output air. Therefore, a system receives zero points if its output air is cooler than 18 C (64.5 F) or warmer than 28 C (82 F). 5.2 Decision Matrix Table 5 shows the decision matrix used to evaluate the different concept variants. The different categories from the objective tree form the columns in the matrix, and the different concept variants form the rows. Each cell in the matrix shows the score given to a particular concept variant in the corresponding category. Table 5. Decision Matrix. Present Worth Humidity Transient Capacity Comfort Total LHS-9000 Desiccant System HCD-9000 EA Desiccant System Vapor- Compression (Water Cooled) Vapor- Compression (Air Cooled) The present worth scores are based on the economic analysis presented in Appendix E. The best scoring concept variant is the water cooled vapor-compression system. The present worth of this system is $582,000. Therefore, this system receives the full 50 points in this category. The other systems were scored accordingly, using the 55

59 present worth values calculated in Appendix E. The desiccant systems are capable of achieving dew points as low as -48 C (-55 F). These two systems receive the maximum points allowed in this category as outlined in Section 5.1. Because vapor-compression systems are feasible, the design team assumes that the vapor-compression systems will meet the specifications, which means that these systems will receive no credit for achieving lower-than-required dew points. The transient response scores are similar in nature to the scores calculated for humidity. The desiccant systems are both able to provide higher-than-specified volumetric flow rates, as is indicated in Table 5. Again, the vapor-compression systems are assumed to exactly meet the prescribed specifications, so they receive no extra points for generating higher flow rates. All of the systems are capable of producing air that is within the human comfort zone, as defined by ASHRAE [ASHRAE, 1989]. Therefore, each of the systems receive the full credit in this category. The decision matrix indicates that the vapor-compression systems rate higher than the desiccant systems, but the matrix does not provide enough resolution to differentiate between the air cooled and the water cooled vapor compression cycles. VI. Final Recommendations The decision matrix presented in the previous section indicates that vaporcompression systems offer the most promising solution to dehumidifying the tunnels inside Mansfield Dam. The vapor-compression systems can achieve the required output 56

60 dew point of 10 C (50 F) for the least cost. The decision matrix, however, indicates no difference between the air cooled and water cooled vapor-compression systems. Differences between these two systems are minimal, and the decision matrix does not have the resolution to differentiate between them. Although the decision matrix cannot discern a difference between the two vaporcompression systems, the design team feels that water cooled vapor-compression system has a slight advantage. The team s calculations leading up to the decision matrix indicate no significant difference between air cooled and water cooled vapor compression systems, but literature indicates otherwise. According to the ASHRAE Handbook, 1979 Equipment, initial and operating costs are typically lower for water cooled systems when an abundant source of water is available. Based on this analysis, the design team recommends that LCRA install a water cooled vapor-compression system to alleviate high humidity levels in Mansfield Dam. VII. Conclusions and Future Work The dominating placed on the design team regarding this design project has been time. Time limitations have limited the development of models, hindered the collection of vendor information, and have even limited the consideration of concept variants. The following paragraphs step through the methodology employed by the team, highlight the weaknesses of the models used and assumptions made in those steps, and recommend 57

61 avenues for future work to improve on the work done by the design team over the course of the semester. 7.1 Redefining Constraints The first area of work that should be reconsidered is the set of constraints identified by the design team. The design team initially constrained the design to utilize the existing tunnel network as the means for air distribution. Although this constraint remains valid, the team made the mistake of subconsciously extending the constraint to require that any dehumidification system had to be located in the existing equipment bunker. This extra constraint limited the flexibility of the solution principles and eliminated the closed loop system from consideration before it was ever conceived. The closed loop system has advantages and disadvantages that must be evaluated as part of the decision making process. Specifically, the closed loop system would probably have a lower initial and operating cost, but disadvantages include obstruction of the tunnels due to the required doors and the likely reduced cooling effect to the power house. 7.2 Inner Surface Temperature Estimates The first parameter listed in the path to system specification section is the temperature of the inner surfaces. The design team estimated the temperatures of the inner surfaces by developing a thermal model of the dam. This model is informative, but the temperatures indicated by it are not as accurate as temperatures that could be directly measured. An accurate profile of the surface temperatures in the various tunnels is 58

62 important to specifying the humidity ratio in the tunnel network. LCRA should consider investing the minimal resources that would be needed to acquire a year-long compilation of this data. 7.3 Evaporation Model Like the other models, the evaporation model is only as accurate as the assumptions upon which it is based. The model is probably accurate enough to support the validity of the calculations that are based on it, but better estimates of the assumed values could improve the accuracy of the model. 7.4 Thermal Model Boundary Conditions The thermal model of the dam can still be used as a useful tool in predicting the temperatures of the tunnel surfaces. One should understand, however, that the accuracy of this model is completely dependent on the accuracy of the boundary conditions. In order for the model to be used with improved accuracy, better data would have to be obtained for the temperatures along each surface of the dam cross section. 7.5 Water Temperature at System Entrance The water temperature at the dehumidification system entrance is a parameter that should be accurately quantified in order to develop a system that is as efficient as possible. The water temperature is important both for determining the air temperature drop through a water cooled heat exchanger and for determining the required high side 59

63 pressure in a water cooled vapor compression system. By accurately determining the maximum possible temperature at which the cooling water may enter the system, a designer can develop a system that is optimized for water temperatures in the proper range. 7.6 Vendor Information The information that the design team was able to obtain from vendors is very incomplete. In order to completely develop the decision matrix, complete vendor information for each type of system must be acquired. The design team estimated missing values as carefully as possible, but an accurate analysis mandates the use of real numbers. The economic analysis shows that the initial costs of each system have little impact on the overall decision. This decision is, however, heavily based on the differences in operating costs among the different systems. This is fortunate, because the design team was able to estimate the operating expenses with more precision than the initial costs. For this reason, the analysis performed by the team should be valid. The team still recommends, however, that a more exhaustive search be made for systems that can meet the specifications developed by the design team. 7.7 Economics The benefits calculated in the economic analysis are completely dependent on the estimated annual savings that the dehumidification system will provide. LCRA should verify the accuracy of the preliminary estimate given to the design team (approximately 60

64 one man-year) before acting on the results of the economic analysis. Concluding Remarks The final recommendation presented in this report is based on a number of imperfect and incomplete models and assumptions. The design team members feel, however, that the final recommendation of a water cooled vapor-compression system with pre-cooling is sound. Because the design team was unable to recommend a specific vendor for the dehumidification system, the most valuable information provided to LCRA in this report is the specification sheet that summarizes the team s extensive analysis of the conditions in the dam that lead to corrosion. Although the final recommendation lacks specific vendor information, the design team has supplied LCRA with an extremely cost effective solution for dealing with the humidity problems in Mansfield Dam. 61

65 References Adams, Debra L., fax concerning desiccant dehumidification system, Low Humidity Systems, Inc., from (770) to Matt Payne at (512) , July 15, Adams, John A. Jr., Damming the Colorado, Texas A&M University Press, College Station, Texas, American Society of Heating, Refrigerating, and Air-Conditioning Engineers, 1989 ASHRAE Handbook, Fundamentals, I-P Edition, American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Atlanta, ATS, Inc., Desicair Desiccant Dehumidifiers, Frederick, Maryland, Barton, Karel, Protection Against Atmospheric Corrosion, Translated: John R. Duncan, John Wiley & Sons, London, Bodine, Grant, interview, telephone conversation with L. C. Eldridge Sales Co. Inc., July 25, Bosich, Joseph F., Corrosion Prevention for Practicing Engineers, Barnes & Noble, New York, Cash, Greg, fax concerning heat exchanger, O & M Manufacturing Co., from (713) to Clark Hughes at (512) , July 30, Chipman, Chip, fax concerning reheat system, Gaumer Company, Inc., from (713) to Clark Hughes at (512) , July 30, Clifford, George, Modern Heating, Ventilating, and Air Conditioning, Prentice Hall, 62

66 Englewood Cliffs, New Jersey, Cross-Sectional View of Mansfield Dam, LCRA handout. "CT, R & RC Series Supplemental Information Performance Specifications & Electrical Data," Sales Brochure, Airflow Company, Maryland. Daugherty, Robert L., and Joseph B. Franzini, Fluid Mechanics With Engineering Applications, 7 th ed., McGraw-Hill, New York, Dehumidification Equipment, Thomas Register, vol. 6, Thomas Publishing Company, New York, Denny, Silina, fax concerning price of desiccant dehumidification system, Low Humidity Systems, Inc., from (770) to Matt Payne at (512) , July 30, Dryomatic, Dehumidification Engineering Manual, Dryomatic Division, Airflow Company, Duncan, Bob, interview, telephone conversation with Tisdale Corporation, July 17, Fohn, Bob, interview, site visit Mansfield Dam, Austin, Texas, June 11, Howell, John R., and Richard O. Buckius, Fundamentals of Engineering Thermodynamics, 2 nd ed., McGraw-Hill, New York, Incropera, Frank P., and David P. DeWitt, Fundamentals of Heat and Mass Transfer, 3 rd ed., John Wiley & Sons, New York, Ingram, Joe Britt, Parametric Performance Study of a Solid Desiccant Rotary Air Dryer, Thesis: The University of Texas at Austin, May James, Ronald W., Desiccants and Humectants, Noyes Data Corporation, Park Ridge, 63

67 New Jersey, Johnson, Mark, concerning lead paint, to July 17, Johnson, Mark, concerning standards hazardous materials, to July 21, Kays, W.P., and A.L. London, Compact Heat Exchangers, 2 nd ed., McGraw-Hill, New York, LaRoche Air Systems Inc., Energy Conservation Wheel, brochure, Baton Rouge, Louisiana, LCRA Blueprint 249-D-833, H. W. Tabor, LCRA Blueprint 249-D-1352, R. E. Glover, LCRA Blueprint 249-D-1364, I. E. Houk, "LCRA Chronicle," published by LCRA, USA, Limnographical Data - Lake Travis, , Defense Research Laboratory, The University of Texas, Austin, Texas, Moreland, Sam, fax concerning power of desiccant dehumidification system, Munters Corp., from (281) to Matt Payne at (512) , July 28, Moreland, Sam, interview, telephone conversation with Munters Corp., July 25, Munters, Dehumidification Handbook, 2 nd ed., Cargocaire, Amesbury, Massachusetts, Munters, HoneyCombe Dehumidifiers,Modular Units, Cargocaire, Amesbury, Massachusetts,

68 Piper, James N., Personal Interview at Applied Research Laboratories, Austin, TX, June 20, Rodak, Ed, fax concerning water cooled dehumidification system, Coolair Corp., from (717) to Stephen Johnson at (512) , July 16, Sumner, Andy, concerning cost/benefit evaluation, to July 14, Sumner, Andy, concerning maintenance costs, to July 9, Sumner, Andy, concerning questions on costs and electricity, to July 17, Sumner, Andy, interview, site visit Mansfield Dam, Austin, Texas, June 11, Trane Co., website, Trane Equipment, date accessed July 30, Williams, John, The Story of the Lower Colorado River Authority, LCRA brochure, published by LCRA,

69 Appendix A: Tail Race Temperatures These temperatures come from LCRA readings of the Mansfield Dam tail race. The tail race of the dam is the body of water directly downstream of the power generation equipment. Water that is in the tail race has entered through the dam by way of three 2.6 m (8.5 ft) diameter pipes that lead to the three turbines in the power house. The penstock inlets are at the same level as the floodgates, 45.7 m (150 ft) below the lake surface. After water has entered the penstocks, the water is directed to the turbines in the power house on the downstream side of the dam. The water drives the turbines, exits the power house, and flows into the tail race. The water in the tail race then flows down the Colorado River. This water absorbs some thermal energy in passing through the dam which raises the temperature readings, but these temperatures are still an approximate reading of water temperature at the level of the penstocks and floodgates. Sometimes the water levels in Lake Travis are quite low. When this happens, as in the drought Central Texas experienced in the summer of 1996, LCRA does not generate power in Mansfield Dam. Because no water is added to the tail race from the turbine outlets, the water in the tail race undergoes thermal stratification. Therefore, the surface temperatures of the tail race are not always a good indication of upstream water temperature because the tail race surfaces are too warm. A1

70 Table A1. Tail race temperature data. Marshall Ford Hydroelectric Facility Tail Race Surface Temperature (degrees F) Monthly Avg. 1-Jan Jan Feb Feb Mar Mar Apr Apr May 56 NA May 59 NA Jun Jun Jul Jul Aug Aug Sep Sep Oct Oct Nov Nov Dec Dec Annual Avg NA: Not Available. This data was obtained from LCRA. A2

71 Appendix B: Mansfield Dam Temperature Distribution Model The design team developed the following Matlab routine to estimate the temperature distribution through a cross-section of Mansfield Dam. This temperature data is necessary because it provides approximate temperatures for each tunnel in the dehumidification air flow network. These temperatures directly affect the evaporation model, as discussed in Appendix D. % tdam.m% The name of the file is tdam.m. clear all % Clears Matlab s memory of any existing variables. % variables % The following is a list of variables to be input into the routine. delta = 4; % meters, the grid size height = 65.9; % meters, height of dam from base to spillway width = 58.8; % meters, width of base of dam (cross-section) topwidth = 10.1; % meters, dam is modeled as a trapezoid - this is the width of the top lakelevel = 63.0; % meters, height of lake surface relative to dam base Tair = 29.44; % degrees C, an average ambient air temperature Tground = 12.8; % degrees C, best estimate for temperature of ground below base of dam epsilon = 0.01; % epsilon convergence criterion numy = round(height/delta) +1; % Calculates number of vertical nodes. numx = round(width/delta) +1; % Calculates number of horizontal nodes. numtop = round(topwidth/delta) + 1; % Calculates number of nodes along the top. slope = -1.0*numy/(numx-numtop); % Calculates slope of the trapezoid. jintercept = (numy*numx)/(numx-numtop); % Calculates y intercept of line formed by back surface of dam. numywater = round(lakelevel/delta)+1; % Calculates number of nodes from base to water surface. TempT = 1; % Initializes a variable called TempT, sets it to 1. iterations = 0; % Initializes a variable called iterations to count number of iterations performed. output = [0,0]; % Initializes a matrix called output. T = zeros(numx,numy); % Initializes a matrix of dimensions numx by numy and sets all elements to 0. for i = 1:numx, % These five lines are a nested for loop that sets all nodes in the matrix T for j = 1:numy, % equal to 15 degrees C. T(i,j) = 15; % The program runs faster this way than if all elements start at 0. end end for i = 1:numx, T(i,numy) = Tair; end % These three lines set the top nodes to Tair. % These five lines set all nodes above the back % surface of the dam to Tair. It loops through % all i s but only calculates for j s whose values % are greater than the value defined by the slope- % intercept form of the equation describing the back % surface. for i = 1:numx, for j = (round(slope*i +jintercept)+1) : numy, T(i,j) = Tair; end end B1

72 for i = 1:numx, T(i,1) = Tground; end for j = 1:numy, T(1,j) = Tair; end % These three lines set the bottom nodes to Tground. % These three lines set all temperatures on upstream surface to Tair. % We will come back to correct the under-water nodes next. % The next few blocks of code take all the under-water nodes that are within 100 feet of the water % surface and match these nodes to a temperature profile that the design team obtained from the % Defense Research Laboratories (now called Applied Research Laboratories) of UT, Austin. % This data provides temperature data for depths up to 150 feet for Lake Travis, averaged over the % months of July and August [Limnographical]. d = [0,5,25,50,100,150]; % Depths (in feet) for temperature profile. % tm = [84.9,84.18,82.44,77.26,56.98,54.36]; % Lake temperature profile in degrees F. tm = [29.39,28.99,28.02,25.14,13.88,12.42]; % Lake temperature profile in degrees C. numdeltas = (30.48/delta); offset = (round(numywater-numdeltas)); % Calculates how many deltas are in 100 feet. % Calculates offset, a variable that tells how many nodes % are below the 100 foot gradient depth. for j = (round(numywater-numdeltas)):numywater, x = ((numywater-j)*delta)*3.2808; increment(j-offset+1) = x; end temporary = interp1(d,tm,increment); tempdisp = (round(numywater-numdeltas))-1; for j = (round(numywater-numdeltas)):numywater, T(1,j) = temporary(j-tempdisp); end % These lines first calculate x values for the points % along the upstream face of the dam, where x is % the distance in feet below the water surface. % This makes the distances consistent with the % temperature profile depths. % Next, the elements in a vector called increment are % set to the x values that correspond to those nodes. % This line builds a vector called temporary that is the % interpolated temperature data for these nodes on the % upstream face. % This line calculates a displacement value to make all the % nodes match up properly. % These three lines set these upstream face nodes % in the matrix T to the proper interpolated values. for j = 1:(round(numywater-numdeltas)), % These three lines set the nodes that are deeper than 100 T(1,j) = 13.1; % feet to a constant 13.1 degrees C, which is the value end % indicated from the temperature profile data for such depths. % Now we solve the interior temperatures numunconverged = 1; while numunconverged > 0, numunconverged = 0; % Initialize a variable called numunconverged and set it equal % to a positive number. % While there are still unconverged nodes, iterate through the % following loop to solve the internal temperatures. % Sets the counter of unconverged nodes equal to zero. for i = 2:(numx-1), % These six lines establish ceiling values that prevent if i <= numtop % the routine from wasting time calculating nodes ceiling = numy-1; % that are outside the dam. else ceiling = (round(slope*i +jintercept)); end for j = 2:ceiling, % Now we set a temporary temperature value for each node equal % to the average of the temperatures of the surrounding nodes. B2

73 TempT = (T(i,(j+1)) + T(i,(j-1)) + T((i-1),j) + T((i+1),j))/4; diff = abs(tempt-t(i,j)); % Establishes a variable called diff to find the change in temperature % of a node from one iteration to the next. if diff > epsilon, % If diff is greater than the epsilon convergence criteria, the number % of unconverged nodes is incremented by one. numunconverged = numunconverged+1; end T(i,j) = TempT; % Having updated numunconverged, the routine replaces the old % temperature of a node with the new calculated value. end % Ends the j looping. end % Ends the I looping. iterations = iterations+1; if (rem(iterations,10) == 0), output(1,1) = iterations; output(1,2) = numunconverged; disp(output) end % Updates the iterations counter. % If iterations is a multiple of 10, then this % block of code displays the number of iterations % as well as the number of unconverged nodes. end % Ends the large while loop. % And now to make the pretty plots clf % Clears the screen of any existing figures. Tplot = zeros(numy,numx); for i = 1:numy, for j = 1:numx, Tplot(i,j) = T(j,i); end end for i=1:numx, X(i,1)=(i*delta); end % Transposes the matrix T and calls the new matrix Tplot. % This makes the plots right-side-up. % These three lines convert deltas to meters and generates a matrix % called X that will be used to generate the x-axis on the plots. for j=1:numy, % Same thing as above, but for Y. Y(j,1)=(j*delta); end hold off % This command makes any subsequent plots a new figure. colormap(jet) % Sets the color scheme: cold = blue, hot = red. pcolor(x,y,tplot) % Plots the temperature data vs. X and Y in what is called a pseudocolor plot. shading interp % Interpolates between elements in Tplot to make the coloration smooth. hold on % Subsequent plots will now be placed on top of the active plot. contour(x,y,tplot,40) % Generates a contour plot with 40 isotherms. axis('square') % Scales the axes so that the plot has the correct aspect ratio. title('temperature Profile Through Dam Cross-Section')% Adds a title. xlabel('distance (meters)') % Adds a label to x axis. ylabel('distance (meters)') % Adds a label to y axis. hold off % Any plots the user generates will be displayed as new figures so that % the current figure is not affected. B3

74 Appendix C: Air Flow Model I. Introduction. The design team modeled air flow patterns in the gate operating galleries to determine what percentage of the air exiting the conditioning system flows through each tunnel. As dry air flows through the tunnels of the dam, it absorbs moisture from wetted surfaces. By the time conditioned air has reached the end of the gate galleries, it must still be dry enough to prevent condensation on cool surfaces. This air flow model allows the team to determine how much conditioned air reaches each part of the gate operating galleries. II. Method Used for Calculations. The team used a method introduced by professor Hardy Cross to generate a spreadsheet that models the flow of air in the dam. The Hardy Cross method uses iteration to determine the flow of a fluid through a pipe network [Daugherty, 1977]. The team s main reference in implementing the Hardy Cross method was Fluid Mechanics with Engineering Applications by Robert L. Daugherty and Joseph B. Franzini, but most standard fluid mechanics textbooks present the Hardy Cross method. The second reference used in generating the air flow model was the American Society of Heating, Refrigerating, and Air Conditioning Engineers (ASHRAE) Handbook, 1989 Fundamentals Volume [ASHRAE, 1989]. Chapter 32 of this handbook C1

75 provides relations for calculating pressure losses through conduits. The ASHRAE Handbook, 1989 Fundamentals also provided minor loss coefficients for flow through elbows and intersections. III. Tunnel Network Schematic. As an initial step in generating the air flow model, the design team created a schematic of the gate operating galleries. These galleries contain the pumps used to raise and lower the flood gates. Figure C1 below presents the schematic of the gate operating galleries. The circled numbers indicate tunnel reference numbers. The team assigned each tunnel a number for ease of reference. Arrows indicate an initial guess of the direction of expected air flow. Shaded boxes indicate the locations of floodgate operating equipment. Dehumidification Bunker 8 Downstream Gates 1-6 Downstream Gates Upstream Gates Upstream Gates 7-24 Figure C1: Airflow through gate operating galleries. C2

76 The first step of the Hardy Cross method is to divide the tunnel network into loops [Daugherty, 1977]. The design team divided the tunnel network into the two loops presented in Figure C2. Loop one houses the equipment to open and close floodgates seven through twenty-four in tunnels two, three, four and five. Loop two house the equipment to operate floodgates one through six in tunnels five, six, seven, and nine. Tunnels one and eight receive all of the air exiting the dehumidification system, so these tunnels are not included in the flow model. Loops one and two share tunnel five. Downstream Gates Downstream Gates Loop Loop Upstream Gates 1-6 Upstream Gates 7-24 Figure C2: Loops used for air flow model. After dividing the tunnel network into loops, the team identified minor losses such as elbows and intersections in the tunnels. Black rectangles in Figure C3 on the following page indicate elements that cause minor pressure losses in the tunnel network. Each element is labeled with a letter. Figure C3 also illustrates the sign convention used in calculating volumetric flow rates. Flows in a counterclockwise direction are considered positive. The arrows inside C3

77 loops one and two denote this sign convention. J K Downstream Gates I H Downstream Gates C B D 6 Loop Loop F 3 L Upstream Gates 1-6 G Upstream Gates 7-24 E Figure C3: Pressure head losses in gate operating galleries. IV. Spreadsheet Calculations. Once the tunnel network was divided into two loops and all of the minor loss elements had been identified, the team began to describe the tunnel network using an Excel spreadsheet. When using the Hardy Cross method, one assumes an initial flow pattern and computes a correction factor for each loop based on pressure losses within the tunnel network [Daugherty, 1977]. The assumed flow rate for each loop is modified by adding this correction factor, and the spreadsheet is recalculated using the new flow rates. This iterative process is repeated until the flow rate correction factors converge to zero. Figure C4 on the following page presents the initial iteration of the team s air flow model spreadsheet. A description of each column of the spreadsheet follows Figure C4. C4

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79 Tunnel #: The first column of the spreadsheet lists each tunnel s reference number. These numbers correspond to the tunnel numbers in Figures C1, C2, and C3. Tunnel Name: The second column lists a short descriptive name for each tunnel. Area: The cross-sectional area for each tunnel is presented in the third column. Figure C5 presents a dimensioned cross-sectional view of a typical gate operating gallery. These dimensions were used for area and perimeter calculations. Perimeter: The fourth column presents the perimeter of each tunnel. D h (hydraulic diameter): Since the tunnels cross-sections are not circular, the hydraulic diameter is used as a characteristic diameter for 1.0 m 2.0 m 3.43 m pipe flow calculations. The hydraulic diameter is defined as follows [ASHRAE, 1989]: Drainage trench Figure C5: Tunnel cross section A D = 4 h P where D h = hydraulic diameter (in) A = tunnel cross-sectional area (in 2 ) P = tunnel perimeter (in) C6

80 Q (volumetric flow rate): For the first iteration of the spreadsheet, a volumetric flow rate is assumed for purposes of initial calculations. The flow rate presented in this column represents the amount of air in cubic feet per second that flows through each tunnel. Counterclockwise flows are considered positive. It is important to note that all calculations to the right of the flow rate (Q) column are based on the flow rate. Flow rate calculations for subsequent iterations are discussed in the section titled Q (flow rate correction factor). V (linear velocity): The linear velocity of the air flowing through each tunnel is presented in the next column. The linear velocity in feet per minute is calculated as follows: V = Q A where V = Linear velocity (feet/min) Q = Volumetric flow rate (ft 3 /min) A = tunnel cross-sectional area (ft 2 ) p v (Velocity Pressure): Velocity pressure is used to calculate the pressure drops in the tunnels. The velocity pressure for air at standard conditions is calculated as follows [ASHRAE, 1989]: V p = v 4005 where p v = Velocity pressure (inches of water) 2 C7

81 V = linear velocity (feet/min) Re (Reynolds number): The Reynolds number is used for pressure loss calculations for each pipe. The Reynolds number for air at standard conditions is calculated as follows [ASHRAE, 1989]: Re = 856. D V h Where Re = Reynolds number D h = Hydraulic diameter of tunnel (inches) V = Linear velocity of flow (feet/min) f (friction factor): The friction factor is used in calculations of pressure drops through straight lengths of tunnel. The friction factor is calculated as follows [ASHRAE, 1989]: where f = friction factor ε f = Re D h if f 0.018, then f if f < 0.018, then f = f = f e = roughness factor, (for concrete e = 0.01 ft) [ASHRAE, 1989] D h = hydraulic diameter (in) Re = Reynolds number C8

82 k1 title, k1 value, k2 title, k2 value: The next four columns present coefficients for the minor losses illustrated in figure C3. These minor loss coefficient were computed using tables in the ASHRAE Handbook 1989 Fundamentals. Each minor loss is assigned a title consisting of a short description, such as elbow, and a letter corresponding to figure C3. The minor loss coefficient is tabulated in the cell to the right of the minor loss coefficient title. Σk (sum of k values): This column sums the k values for each tunnel. These values are used in pressure loss calculations and are computed for each tunnel as follows: k = k1 value + k2 value P (pressure drop): The pressure drop through a tunnel is calculated using the Darcy-Weisbach Equation [ASHRAE, 1989]: 12 fl P = + D h k p where f = friction factor L = tunnel length (feet) D h = hydraulic diameter (inches) Σk = sum of minor loss coefficients v C9

83 p v = velocity pressure (inches of water) Q (flow rate correction factor): The actual flow rate through the tunnel may be assumed to be a function of the assumed flow rate and a correction factor [Daugherty, 1977]: where Q= Q0 + Q Q = actual flow rate (ft 3 /min) Q 0 = assumed flow rate (ft 3 /min) Q = flow rate correction factor (ft 3 /min) The equation for pressure drop through a tunnel may also be written as follows [Daugherty, 1977]: where P= KQ n K = loss coefficient for each tunnel Q = flow rate through the tunnel n = flow rate exponent. For the Darcy-Weisbach Equation, n = 2. Expanding P = KQ n using Q = Q 0 + Q and the binomial expansion theorem yields the following [Daugherty, 1977]: n n n 1 ( 0 ) ( o 0...) n P= KQ = K Q + Q = K Q + nq Q+ C10

84 If Q is small in comparison to Q 0, the rest of the terms in the binomial expansion may be neglected. Solving the above equation for Q yields the following [Daugherty, 1977]: n KQ Q = = 0 n 1 KnQ0 n or P Q = n P Q 0 P P Q 0 where P = pressure drop through a tunnel (inches of water) Q 0 = assumed flow rate used to calculate Q correction factor n = flow rate exponent (n=2 for Darcy-Weisbach Equation) Once Q has been calculated for one iteration, the assumed flow rate for the next iteration is calculated as follows: where Q= Q0 + Q Q = flow rate for next iteration (ft 3 /min) Q 0 = assumed flow rate for current iteration (ft 3 /min) Q = flow rate correction factor (ft 3 /min) A new iteration is then calculated using the new assumed flow rate. This process C11

85 is repeated until the flow rate correction factor ( Q) approaches zero. The flow rates yielded by the team s air flow model are presented in figure C6 below. Each tunnel s air flow rate in cubic feet per minute is listed next to its tunnel number. These flow rate values were used with the evaporation model to determine the required performance of the dehumidification system. Figures C7 through C11 on the following pages present iterations one, two, three, thirty-three and thirty-four. For purposes of the flow model, the design team s final iteration was iteration thirty-four. Dehumidification Bunker 9000 cfm 8 Downstream Gates 1-6 Downstream Gates cfm cfm cfm cfm cfm cfm cfm cfm Upstream Gates 1-6 Upstream Gates 7-24 Figure C6: Air flow model results. References American Society of Heating, Refrigerating, and Air-Conditioning Engineers, 1989 ASHRAE Handbook, Fundamentals, I-P Edition, American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Atlanta, Daugherty, Robert L., and Joseph B. Franzini, Fluid Mechanics with Engineering C12

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91 Appendix D: Evaporation Model The team developed this evaporation model in order to estimate the humidity ratio at all points in each tunnel of interest. This model serves as a tool for correlating minimally acceptable tunnel conditions to the input conditions for any particular tunnel. The following analysis outlines the development of this model and demonstrates its use in determining required tunnel input conditions. Evaporation is a mass transfer phenomenon that is analogous to heat transfer. The transfer of water from the tunnel surfaces to the moving air is driven by a difference in species concentrations between the saturated air at the tunnel surface and the drier air flowing closer to the center of the tunnel. This mass transfer is quantified by the following equation [Incropera, 1996]. where: n = h ( ρ ρ ) '' s m s m n s '' is the mass flux rate per unit area at the surface. h m is the mass transfer coefficient. ρ s and ρ m are the surface and mean species concentrations, respectively. The mass transfer coefficient is calculated from the Sherwood number, Sh D, as follows [Incropera, 1996]. Sh D hmd or h D AB m ShDD = D AB D1

92 where: D AB is the binary diffusion coefficient for water vapor and air, m 2 /s [Incropera, 1996]. D is the hydraulic diameter of the tunnel, D Area = 4. Perimeter The next step is to calculate the Sherwood number [Incropera, 1996]. Sh D = Re 45 / D Sc n where: Re D is the Reynolds number, Re VD ν, where V is the mean air velocity, D is the tunnel hydraulic diameter, and ν is the air kinematic viscosity [Fox, 1992]. Sc is the Schmidt number, Sc n is approximately ν [Incropera, 1996]. D AB Figure D1 shows, in a graphical form, how these different parameters are used to determine the local mass flux rate. 1 In the analogous heat transfer correlation, n is 0.3 for cooling and 0.4 for heating [Incropera, 1996]. The design team chose 0.35 as a best approximation. Note that this exponent has little effect on Sh for values of Sc near 1. D2

93 D AB ν D V Sc Re D AB D Sh h m ρ ρ s m n s '' Figure D1. Path to calculating the mass flux rate. The parameter n s '' indicates the evaporation rate at any given point along a tunnel. This evaporation causes the species concentration of the water vapor in the air to increase, which results in a decrease in ρ s ρ m. This reduction of the evaporation driving force causes the evaporation rate to drop as a function of distance through any particular tunnel. Due to the variable nature of this evaporation rate, the design team chose to use an iterative method to accurately quantify the change in humidity through each tunnel. The file alltunls.xls is included on a floppy diskette with this document. The spreadsheet is very easy to use, but the following paragraphs are included to assist the unfamiliar user. The first sheet is labeled input-output. This sheet, shown in Figure D2, provides humidity ratio data to the other sheets. The input humidity ratio value for tunnel D3

94 number one is the same as the output humidity ratio for the dehumidification system that is installed. The user can change this value to evaluate the effect of this parameter on the humidity ratio throughout the treated tunnels. The next column of data is the saturation humidity ratio values for each tunnel. These values are the humidity ratios that correspond to saturation conditions at the temperature of the surface of each tunnel, as determined from a psychrometric chart. For example, a tunnel with a 15 C surface temperature would have a saturation humidity ratio of kg water /kg air. These values are automatically fed into the other sheets in the file and used in the evaporation calculations. The design team performed a thermal analysis of Mansfield Dam to determine the tunnel temperatures as accurately as possible. It is not recommended that these numbers be changed unless the user has access to accurate tunnel surface temperature data throughout the dam. Figure D2. Section of input-output spreadsheet. D4

95 The sheet labeled results is a summary of the results generated by the air flow model presented in Appendix C. The user may alter columns I and J, shown in Figure D3, to reflect the output from the air flow model if parameters within that model are changed. The spreadsheets labeled tunnel 1 through tunnel 9 calculate the changing humidity ratio through each of the nine tunnel segments. The only parameter that the user would conceivably change in these sheets is the wetted surface area per Figure D3. Columns I and J from sheet results of the spreadsheet named alltunls.xls. meter length. This value currently reflects the width of the water-filled trenches in each tunnel, but it could be changed to include areas that are wet due to seepage through expansion joints. The user may also increase this value to represent the current wetted surface area in the tunnels, which is largely comprised of condensate-coated surfaces. These surfaces will no longer be wet after a dehumidification system has been operating for some time, but the user may wish to alter this value to evaluate the transient drying time of such a system. The results of all the calculations performed by this spreadsheet are summarized in the sheet labeled tot length. This spreadsheet lists the changing humidity ratio through each tunnel in the upstream circuit, the circuit formed by the tunnels labeled 1, 2, 3, 6, 9, and 8 in Figure C1. This circuit is important because it includes the upstream floodgate gallery, the tunnel that currently exhibits the D5