ANALYSIS AND DESIGN OF DESICCANT COOLING SYSTEMS by Kamal A. Abou-Khamis Submitted in Partial Fulfillment ofthe Requirements for the degree of Master ofScience in Engineering in the Mechanical Engineering Program YOUNGSTOWN STATE UNIVERSITY 06,2000 ANALYSIS AND DESIGN OF DESICCANT COOLING SYSTEMS by Kamal A. Abou-Khamis I hereby release this thesis to the public. I understand this thesis will be housed at the Circulation Desk ofthe University Library and will be available for public access. I also authorize the University ofother individuals to make copies ofthis thesis as needed for scholarly research. Signature: KamaL.l\.. Abou-Khamls, Student Peter 1. Kasvinsky, Dean ofGradua 6-~-OO Date Date Date Date 6-~-OO ABSTRACT In the last decade, desiccant dehumidification has emerged as an alternative or as a supplement to conventional vapor compression systems for cooling and conditioning air in commercial and industrial buildings. It provides a method of drying air before it enters a conditioned space. When combined with conventional vapor compression systems, desiccant dehumidification systems are a cost-effective means of supplying cool, dry, filtered air. The object of this study is to investigate the performance of a desiccant cooling unit for a large building supermarket in Canfield, Ohio. Optimization of this system is done by computer simulation. Then a conventional vapor compression system is designed for the supermarket and the performances of the two systems are compared. This comparison shows that the desiccant unit performs better than a conventional unit in the environment of the supermarket. Another case study is to design and investigate the performance of a desiccant cooling unit in a more humid area, Tampa, Florida. This case study shows that the desiccant performs better in regions with large specific humidity. A further study deals with the design improvement on the desiccant units to enhance the units' performance. The improvement focuses on pre-cooling the make-up air and dehumidifying it with the desiccant before the air blends with return air from the zone. The study shows that significant improvement can be made for the cooling units used in Florida. 111 ACKNO\VLEDEGMENTS First and foremost, a great deal of gratitude goes to Dr. H.W. Shawn Kim who served as my thesis advisor. His patience and willingness to teach me about Heat Transfer and Air-Conditioning Systems are immensely appreciated. Dr. Ganesh V. Kudav and Dr. Robert A. McCoy served on my thesis committee, and their feedback on my work was quite valuable. Finally, Michael Palotsee has provided me with a great deal of information and advice throughout my thesis, and also deserve my gratitude. My thanks go to all ofthese individuals. Also, my appreciation goes to my brother and friend Mohammad Abou-Khamis, for his help and support. My ultimate gratefulness is for my parents and the rest of my family, who have enthusiastically supported all of my academic undertakings. I share this accomplishment with them all. IV TABLE OF CONTENTS PAGE CHAPTER I: STATEMENT OF THE PROBLEM 1 CHAPTER II: THEORY OF DESICCANTS AND DESICCANT COOLING SYSTEMS 3 2.1 Desiccants 4 2.2 Types of Desiccants 6 2.3 Desiccant Life 8 2.4 The Desiccant Cycle 9 2.5 Desiccant Applications 11 2.6 Desiccant Cooling System 13 CHAPTER III: ANALYSIS OF THE COOLING SYSTEMS 16 3.1 Domain of the systems 16 3.2 Computer Simulations 19 3.3 Systems Analysis 22 3.4 Comparison of the two Systems 42 3.5 Recommendations for future choosing between Desiccant and Conventional units 42 CHAPTER IV: SYSTEM DESIGN IMPROVEMENT 45 4.1 Additional Study on Desiccant Systems in Humid Areas 45 4.2 Comparison between the desiccant units in Ohio and Florida 48 4.3 Parameters affecting System Perfonnance 52 4.4 Alternative Designs for Improvement 55 4.5 Comparison between the systems 65 CHAPTER V: RESULTS AND CONCLUSIONS 68 5.1 Results 68 5.2 Conclusions 70 BIBLOGRAPHY 72 APPENDIX 74 v LIST OF TABLES TABLE TITLE PAGE 3.1 Design Conditions- Main Sales Area 23 3.2 Design Airflow Schedule/ Airflow Configuration 24 3.3 System Data Values for the Desiccant System 26 3.4 COP for Desiccant, OH 30 3.5 Design Weather Parameters for Youngstown, OH 34 3.6 Space Input Data for the Giant Eagle Store 35 3.7 System Psychrometric for the Conventional Unit 37 3.8 COP for the Conventional Unit 40 4.1 System Data Values for the Desiccant System, FL 47 4.2 COP for Desiccant, FL 49 4.3 Air System Design Load Summary 53 4.4 System Data Values for the Design System, OR 57 4.5 COP for the New Design System, OH 59 4.6 System Data Values for the Design System, FL 62 4.7 COP for the New Design System, FL 63 AJ A.2 Cooling Design Temperature Profiles Air System Sizing Summary for the Conventional Unit VI 76 78 LIST OF FIGURES FIGURE TITLE 2.1 The Desiccant Cycle 2.2 Desiccant Cooling System 2.3 Psychrometric Chart PAGE 10 14 14 3.1 Temperature Distribution in left side ?f the store 18 3.2 Desiccant Cooling Unit, OR 25 3.3 Energy required to run the Desiccant System, OR 31 3.4 Conventional Cooling Unit 36 3.5 Psychrometric Analysis for the Conventional Unit 38 3.6 Energy required to run the Conventional System 41 3.7 Energy required to run the Desiccant and Conventional Systems, OR 43 4.1 Desiccant Cooling System, FL 46 4.2 Energy required to run the Desiccant System, FL 50 4.3 Energy required to run the Desiccant Systems in OH and FL 51 4.4 Changing Process Air Moisture 54 4.5 Changing Process Air Temperature 54 4.6 Alternative System Design, OH 56 4.7 Energy required to run the Alternative System, OH 60 VII LIST OF FIGURES (CONT'l)) FIGURE TITLE PAGE 4.8 Alternative System Design, FL 61 4.9 Energy required to run the Alternative System, FL 64 4.10 Energy required to run the Desiccant, OH and its Alternative Design 66 4.11 Energy required to run the Desiccant, FL and its Alternative Design 67 5.1 Coefficient of Performance 69 A.l Picture of a Desiccant Cooling System 74 A.2 Process and reactivation airflow temperature and humidity changes 75 A.3 Design Temperature Profile 77 VIII CHAPTER I STATEMENT OF THE PROBLEM Large commercial buildings, such as supermarkets or restaurants, require large capacity air conditioning units to maintain a comfortable environment for their customers and employees. These air conditioning units control not only the indoor temperature but also the moisture that may be generated by people, cooking, and other processes. The installation and operating costs of these units are usually high. Many such buildings, in recent years, are being equipped with desiccant cooling units that reduce the indoor moisture level with an expectation that these relatively new devices will increase the performance of the overall air-conditioning system, and thus reduce the unit capacity and the operation costs. However, the performance assessment of desiccant system have not been extensively reported or published. The purpose of this thesis is to investigate the performance of the air conditioning system for a large supermarket building and to find an alternative design for improvement of the system. A portion of the supermarket is considered for this study which contains an open-faced multi-deck meat, multi-deck dairy, produce and frozen food cases. The air conditioning system must supply treated air to the main sales area of the building. The supermarket requires additional latent cooling (dehumidification) to prevent condensation 1 on frozen food. The only option available in a conventional air conditioning system is to tum down the thennostat, resulting in both latent and sensible cooling. This leads to air temperatures that are uncomfortably cold for the customersc Gas fired desiccant dehumidification breaks the link between sensible and latent cooling, allowing building operators to precisely control each to its required level. In this study, the data relevant for routine operating conditions are shown in Table 3.2. These data were used to study and analyze the desiccant cooling unit. A conventional cooling unit was also designed for the same area in order to compare the coefficient of perfonnance for the two units. 2 CHAPTER II THEORY OF DESICCANTS AND DESICCANT COOLING SYSTEMS Desiccants exhibit an affinity so strong for moisture that they can draw water vapor directly from the surrounding air. This affinity can be regenerated repeatedly by applying heat to the desiccant material to drive off the collected moisture. Desiccants are placed in dehumidifiers, which have traditionally been used in tandem with mechanical refrigeration in specialty air-conditioning systems. The systems have been more commonly applied m typical air-conditioning situations that involve large dehumidification load fractions. This situation often arises due to low humidity levels required for operations in many industries. Lower humidity levels, below the level necessary for comfort, are generally unattainable cost-effectively with mechanical refrigeration and reheat. Desiccant dehumidification technology has been used m military storage and many industrial applications for more than 60 years. Continuous desiccant dehumidification can be achieved in a number of ways using liquid spray tower, solid packed tower, rotating horizontal bed, multiple vertical bed, and rotating wheel. In the past, desiccant dehumidification was also integrated with mechanical refrigeration in early air-conditioning approaches for comfort in business and homes. To the present day, though, in specialty applications, desiccant dehumidification 3 still holds its economic advantage over dehumidification by mechanical refrigeration and reheat. In industrial air-conditioning, numerous moisture sensitive manufacturing and storage applications utilize desiccant dehumidification. The dramatic decrease in product rejection, and thus the direct increase in profitability of the product, yields a quick payback on the initial investment in depressed humidity control equipment. Interest is now being revived in thermal-driven desiccant dehumidification in non industrial air-conditioning applications to offset rising electricity prices. Lower cost thermal energy, including natural gas, waste heat, solar energy, and other sources, is substituted for electric energy to meet the dehumidification load on the air-conditioning system. In the past, available desiccant dehumidification equipment has been considered too expensive, compared to assembly-line mechanical refrigeration equipment, for application outside the industrial field of use. But today, desiccant dehumidification technology, supported by ongoing research and development, is providing a cost-saving to reduce electric air-conditioning capacity and thus to lower electric-energy costs and power demand charges in certain nonindustrial air conditioning situations too. 2.1. Desiccants Many materials are desiccants; that is they attract and hold water vapor. Wood, natural fibers, clays, and many synthetics attract and release moisture like commercial desiccants do, but they lack the holding capacity of some special desiccant materials. For example, woolen carpet fibers attract up to 23 % of their dry weight in water vapor, and nylon can take up almost 6 % of its weight in water. In contrast, a commercial desiccant takes up between 10 and 1100% of its dry weight in water vapor, depending on its type 4 and the moisture available in the environment [I]. Furthermore, commercial desiccants continue to attract moisture even when the surrounding air is relatively dry, a characteristic that other materials do not share. All desiccants behave in a similar way in that they attract moisture until they reach equilibrium with the surrounding air. Moisture is usually removed from the desiccant by heating it to temperatures between 120 and 500 of and exposing it to a scavenger airstream. After the desiccant dries, it must be cooled so it can attract moisture once again. Sorption refers to the binding ofone substance to another. It always generates sensible heat equal to the latent heat of water vapor taken up by the desiccant, plus an additional heat of sorption that varies between 5 and 25 % of the latent heat of the water vapor. This heat is transferred to the desiccant and the surrounding air. The process ofattracting and holding moisture is described as either adsorption or absorption, depending on whether the desiccant undergoes a chemical change as it takes on moisture. Adsorption does not change the desiccant except by the addition of the weight of water vapor, similar in some ways to a sponge soaking up water. Absorption, on the other hand, changes the desiccant. An example of this is table salt, which changes from a solid to a liquid as it absorbs moisture. Sorbents are materials that have an ability to attract and hold gases or liquids. They can be used to attract gases or liquids other than water vapor, a characteristic that makes them very useful in chemical separation processes. Desiccants are subset of sorbents; they have a particular affinity for water. 5 2.2. Types Of Desiccants Desiccants can be solids or liquids and can hold moisture through adsorption or absorption. Most absorbents are liquids, and most adsorbents are solids [I]. 2.2.1. Liquid Absorbents Liquid absorption dehumidification can best be illustrated by comparing it to the operation of an air washer. When air passes through an air washer, its dewpoint approaches that of the temperature of the water supplied to the machine. More humid air is dehumidified and less humid air is humidified. In a similar manner, a liquid absorption dehumidifier contacts air with a liquid desiccant solution. The liquid has a vapor pressure lower than water at the same temperature, and when the air passing over the solution approaches this reduced vapor pressure, then it is dehumidified. The vapor pressure of a liquid absorption solution is directly proportional to its temperature and inversely proportional to its concentration. In standard practice, the behavior of a liquid desiccant can be controlled by adjusting its concentration, its temperature, or both. Desiccant temperature is controlled by simple heaters and coolers. Concentration is controlled by heating the desiccant to drive moisture out into a waste airstream or directly to the ambient. As a practical matter, however, the absorption process is limited by the surface area ofa desiccant exposed to the air being dehumidified and the contact time allowed for the reaction. More surface area and more contact time allows the desiccant to approach its theoretical capacity. Commercial desiccant systems reflect these realities either by 6 spraying the desiccant onto an extended surface much like in a cooling tower, or holding a solution in a rotating extended surface with a large solution capacity. 2.2.2. Solid Adsorbents Adsorbents are solid materials with a tremendous internal surface area per unit of mass; a single gram can have more than 50,000 ft 2 of surface area. Structurally, they resemble a rigid sponge, and the surface of the sponge in turn resembles the ocean coastline of a fjord. This analogy indicates the scale of the different surfaces in an adsorbent. The fjords can be compared to the capillaries in the adsorbent. The spaces between the grains of sand on the fjord beaches can be compared to the spaces between the individual molecules of the adsorbent, all of which have the capacity to hold water molecules. The bulk of the adsorbed water is contained by condensation into the capillaries, and the majority ofthe surface area that attracts individual water molecules is in the crystalline structure ofthe material itself [1]. Adsorbents attract moisture because ofthe electrical field at the desiccant surface. The field is not uniform in either force or charge, so it attracts polarized water molecules that have an opposite charge from specific sites on the desiccant surface. When the complete surface is covered, the adsorbent can hold still more moisture, as vapor condenses into the first water layer and fills the capillaries throughout the material. As with liquid absorbents, the ability ofan adsorbent to attract moisture depends on how much water is on its surface compared to how much water is in the air, That difference is reflected in the vapor pressure at the surface and in the air. The adsorption behavior of solid adsorbents depends on (l) their total surface area, (2) the total volume 7 of their capillaries, and (3) the range of their capillary diameters. A large surface area gives the adsorbent a larger capacity at low relative humidities. Large capillaries provide a high capacity for condensed water, which gives the adsorbent a higher capacity at high relative humidities. A narrow range of capillary diameters makes an adsorbent more selective in the vapor molecules it can attract and hold; thus, some will fit and others will be too large to pass through the passages in the material. 2.3. Desiccant Life The useful life of desiccant materials depends largely on the quantity and type of contamination in the airstreams they dry. In commercial equipment, desiccants last between 10,000 and 100,000 hours and longer before they need replacement [1]. Normally, two mechanisms cause the loss of desiccant capacity: change in desiccant sorption characteristics through reactions with contaminants, and loss ofeffective surface area through clogging or hydrothermal degradation. Liquid absorbents are more susceptible to chemical reaction with airstream contaminants other than water vapor than are solid adsorbents. For example, certain sulfur compounds can react with lithium chloride to form lithium sulfate, which is not a desiccant. If the concentration of such compounds in the airstream were below 10 ppm and the desiccant were in use 24 hours a day, the capacity reduction would be true; this may mean a 10 % reduction in capacity over the course ofa year. In air-conditioning applications, desiccant equipment is designed to minimize the need for desiccant replacement in much the same way that vapor compression cooling systems are designed to avoid the need for compressor replacement Unlike filters, 8 desiccants are seldom intended to be frequently replaced during normal service in an air drying application. 2.4. The Desiccant Cycle All desiccants function by the same mechanism-transferring moisture because ofa difference between the water vapor pressure at their surface and that of the surrounding air. When the vapor pressure at the desiccant surface is lower than that of the air, the desiccant attracts moisture. When the surface vapor pressure is higher than that of the surrounding air, the desiccant releases moisture. Figure 2.1 shows the relationship between the moisture content ofthe desiccant and its surface vapor pressure. As the moisture content of the desiccant rises, so does the water vapor pressure at its surface. At some point, the vapor pressure at the desiccant is the same as that of the air and the two are in equilibrium. Then moisture cannot move in either direction until some external force changes the vapor pressure at the desiccant or in the air. Figure 2.1 also shows the impact of temperature on the vapor pressure at the desiccant. Both higher temperatures and increased moisture content increase the vapor pressure at the surface. When the surface vapor pressure exceeds that ofthe surrounding air, moisture leaves the desiccants process called reactivation or regeneration. After the desiccant is dried (reactivated) by the heat, its vapor pressure remains high, so that it has very little ability to absorb moisture. Cooling the desiccant reduces its surface vapor pressure so it can absorb moisture once again. The complete cycle is illustrated in Fig 2.1. The operating economics ofdesiccants depends on the energy cost ofmoving a 9 250 0 '- o Q.Q) ~~ '-- Q) ' _ :::J ca(/) ~c: 0>(\1 c: <.> .- <.> en .- ca en Q) Q) '-0 <.> c: _ ca 0> C o H .~ en Q) o 50 0 Increasing Desiccant~ Moisture Content Fig. 2.1 The Desiccant Cycle 10 ~c: .~ given material through this cycle. The dehumidification ofair (loading the desiccant with water vapor) generally proceeds without energy input, other than fan and pump costs. The major portion ofenergy is invested in regenerating the desiccant (moving from point 2 to point 3) and cooling the desiccant (point 3 to point 1). Regeneration energy is equal to the sum ofthree variables [1]: 1. The heat necessary to raise the desiccant to a temperature high enough to make its surface vapor pressure higher than the surrounding air. 2. The heat necessary to vaporize the moisture it contains (1060 Btullb). 3. The small amount ofheat from desorption ofthe water from the desiccant. The cooling energy is proportional to the mass ofthe desiccant, and the difference between its temperature after regeneration and the lower temperature that allows the desiccant to remove water from the airstream once again. The cycle is similar when desiccants are regenerated using pressure differences in a compressed air application. The desiccant is saturated in a high-pressure chamber, i.e., that of the compressed air. Then valves open, isolating the compressed air from the material, and the desiccant is exposed to air at ambient pressure. The vapor pressure of the saturated desiccant is much higher than ambient air at normal pressures, so the moisture leaves the desiccant for the surrounding air. An alternate desorption strategy uses a small portion ofthe dried air, returning it to the moist desiccant bed to reabsorb the moisture, then venting the air to the atmosphere at ambient pressures. 2.5. Desiccant Applications Desiccants can dry either liquids or gases, including ambient air, and are used in many 11 air-conditioning applications, particularly when [1]: l) The latent load is large in comparison to the sensible load. 2) The cost of energy to regenerate the desiccant is low when compared with the cost of energy to dehumidify the air by chilling it below its dewpoint. 3) The moisture control level required in the space would require chilling the air to subfreezing dewpoints ifcompression refrigeration alone were used to dehumidify the alr. 4) The temperature control level required by the space or process requires continuous delivery ofair at subfreezing temperatures. In any ofthese situations, the cost ofrunning a vapor compression cooling system can be very high. A desiccant process may offer considerable advantages in energy, the initial cost of equipment, and maintenance. Since desiccants are able to absorb more than simply water vapor, they can remove contaminants from airstreams to improve indoor air quality. Desiccants have been used to remove organic vapors, and in special circumstances, to control microbiological contaminants. Desiccants are also used in drying compressed air to low dewpoints. In this application, moisture can be removed from the desiccant without heat. Desorption is accomplished using differences in vapor pressures compared to the total pressures ofthe compressed and ambient pressure airstreams. Finally, desiccants are used to dry the refrigerant circulating in air- conditioning and refrigeration systems. This reduces corrosion in refrigerant piping and avoids valves and capillaries becoming clogged with ice crystals. In this application the desiccant is not regenerated; it is discarded when it has adsorbed its limit ofwater vapor. 12 2.6. Desiccant Cooling System A typical desiccant cooling system consists of a desiccant wheel, a heat exchanger, two evaporative coolers and associated blowers for air movement as seen in Fig. 2.2 After the desiccant wheel adsorbs moisture from the process air, this air exits the wheel hot and dry, due to the desiccant's heat of adsorption. For an effective cooling system, all or most of this heat must be rejected. This is typically accomplished using an air-to-air stationary heat exchanger or a rotary heat wheel. The cooler, dry air leaving the heat exchanger is then passed through an evaporative cooler, which adds moisture to the air, reducing its temperature before it enters the conditioned space [12]. On the regeneration side, air is first reduced in temperature by passing it through an evaporative cooler. This cooled air provides a heat sink for the air-to-air heat exchanger. The hottest air exiting the heat exchanger is used for regeneration of the desiccant wheel. The remainder, if any, is normally rejected outside. The regeneration heat source for the desiccant can be the condenser coil of a boiler system, a direct-fired burner or a waste-heat source. The hot air exiting the heat source is passed through the regeneration section of the desiccant wheel, and the moisture released by the desiccant is rejected to the flue [12]. Desiccant cooling systems can be operated either in a recirculation mode or a ventilation mode. In a recirculation mode, the process inlet air is the return air from the building and the regeneration inlet air is outdoor air. In a ventilation mode, the process inlet air is outdoor air and the regeneration inlet air is either outdoor air (standard vent cycle) or building exhaust air (Pennington cycle) [12]. The state points ofa typical desiccant cooling process operating in a recirculation mode are shown in Fig. 2.3. Ifthe 13 Evo.por-o. tive Coolers Dut:side Bl--irner Flue P"'-'ocess Inlet Fig. 2.2 Desiccant Cooling System 0.028 ... 'w 0.024 ?: '0 "0 0.02 § 0.016 f :;, lil 0.012 "0 E 0.008 is c 5 0.004 0. / ./ / / .' 5040 H~t . Excha tor ==:':"-~--'--~-L-_~---l.'::"--J-...........JO 60 70 80 90 100 110 120 130 Dry Bulb Temperature. OF 30 Fig. 2.3 Psychrometric Chart 14 H~t ==:':"-~--'--~-L-_~---l.'::"--J-...........JO system operates effectively, air entering the cooling system at ARl indoor condition, 80 0 F/ 50% relative humidity (RR), can be delivered to a building mostly saturated at 56 58 0 F. Alternatively, the degree of saturation achieved in the evaporative cooler can be reduced and the air will enter the building slightly warmer but lower in humidity. Desiccant wheels used for cooling applications are designed with channels which provide laminar flow, maintaining the heat and mass transfer area as high as possible with a minimal pressure drop. Process and regeneration sections of the wheel are sealed to minimize cross-leakage. To achieve a small wheel size and optimum wheel efficiency, the following characteristics are desirable in the wheel [12]: 1. High surface area to volume ratio for fast heat and mass transfer. 2. High-temperature regeneration to minimize the size ofthe regeneration section. 3. Desiccant having a Type 1 M isotherm shape to enhance the containment ofmoisture wavefronts on adsorption and the temperature wavefront on regeneration 4. High desiccant/substrate loading ratio. 5. Desiccant having a low heat ofadsorption or a high percentage ofits capacity within a low temperature range. 6. Desiccant able to withstand direct-fired regeneration to eliminate losses in efficiency from a boiler 7. Desiccant properties which are stable over the projected wheel life Characteristics listed in 6 and 7 above relate to the adsorption stability of the desiccant with time and exposure to combustion products. The performances obtained from a desiccant cooling system are obviously directly related to the intrinsic properties of the desiccant. 15 CHAPTER III ANALYSIS OF THE COOLING SYSTEMS 3.1. Domain of the systems The area of analysis is the main sales area of the Giant Eagle Store in Canfield, OHIO. This sales area contains open-faced multi-deck meat, multi-deck dairy, produce and frozen food cases. There are two types of open, refrigerated, display cabinet: one with a condenser at the bottom of the cabinet and the other with a remote condenser, outside the conditioned space. In the former all the power used by the compressors is an extra load on the room and there is no benefit from the heat absorbed by the frozen food in the cabinets themselves. Stores with this type of display cabinet do not suffer from the underheating sometimes experienced by those having the other type. The second type of open refrigerated cabinet, which is a subject of this study, has a big impact on the air conditioning load. Heat transferred to the cabinets comes from the conditioned space and is therefore reducing the sensible heat gain because it is ultimately rejected at the remote condensers. This effect, plus the latent cooling also done at the cabinets, is very significant and must be taken into account when calculating the heat gains, the cooling load and the ratio of sensible to total heat in the central air-handling plant. Refrigerated cabinets with remote condensers remove heat from the sales area over 24 hours of the day and 365 days of the year, Therefore, regardless of the room 16 temperature, underheating is sometimes a difficulty at unexpected times, Fig. 3.1, which was prepared using 'Encore 2100' software, shows a picture of the area concerned, the positions of the cabinets, the frozen food cases, and the temperatures that must be maintained inside them. 17 Bb 1\.1EJ\T COOLE~ IEI!l!m All) (a,.c Color" f;_.~,~~~~C.",~ :: ,': ; ,t ~1l5<'rIlellcllllgs ,F¥ '< - : ,;, ; 'J ",'~ r'vran .. 1 rn.uS'... Uf1It ini)ln~n <. .; Yell....,-BDdS.....,. Fig. 3.1 Temperature Distrubution in the left side of the store 18 COOLE~ f;_.~,~~~~C.",~ ~1l5<'r ~ i)ln~n 3.2. Computer Simulations 3.2.1. Desiccant Unit Selection This Desiccant Unit Selection software is designed to choose the suitable desiccant unit required for any system. The desiccant units are designed from model DC 020, DC 030, to DC 130 and each unit is chosen according to the amount of air entering to the desiccant wheel. This software is not a tool for estimating cooling loads, so the engineer must use another software to find the loads for the building and, then, use the results as an input to the Desiccant Software. Input data include the ambient and return air and the design conditions of the zone. Engineers can choose the components ofthe system to achieve the required results. The Desiccant System is composed ofthese components: 1. Desiccant rotor 2. Heat exchanger rotor 3. Hydronic heater/hydronic loop 4. Regeneration heating coil 5. Process heating coil (optional) 6. Electrical control system 7. Process air blower 8. Regeneration air blower 9. Evaporative pad (optional) 10. Pre-cooling coil (optional) 1L Post-cooling coil (optional) 19 3.2.2. Hourly Analysis Program (HAP) HAP is a computer tool that assists engmeers In designing HVAC systems for commercial buildings. It is a tool for estimating loads and designing systems. HAP uses the ASHRAE-endorsed transfer function method for load calculations. • HAP System Design Features HAP estimates design cooling and heating loads for commercial buildings in order to determine required sizes for HVAC system components. Ultimately, the program provides information needed for selecting and specifying equipment. Specifically, the program performs the following tasks: 1. Calculates design cooling and heating loads for spaces, zones, and coils in the HVAC system. 2. Determines required airflow rates for spaces, zones and the system. 3. Sizes cooling and heating coils. 4. Sizes air circulation fans. • Using Hap to design systems and plants This section briefly describes how to use the HAP to design systems and plants in conceptual terms. All design work requires the same general five-step procedure: 1. Define the Problem; the scope and objectives ofthe design analysis are defined. 2. Gather Data; before design calculations can be performed, information about the building, its environment and its HVAC equipment must be gathered. This step involves extracting data from building plans, evaluating building usage and studying HVAC system needs. 3. Enter Data into HAP; data for climate, building and HV AC equipment must be 20 entered. When usmg HAP, the mam program window allows easy access to the operation. 4. Use HAP to Generate Design Reports. Once weather, space, air system and plant data has been entered, HAP can be used to generate system and plant design reports. 3.2.3. Encore 2100 The Encore 2100 is able to monitor and control all HVAC and utility functions in the largest stores. The Encore 2100 will: • Monitor and control 256 input and 256 output devices • Call personally by phone when an alarm occurs, even when the store is closed • Provide historical information on system performance, using graphs and data When the system is turned on, the main screen will appear. The entire program will be run from this screen by use ofthe function keys in the menu at the bottom ofeach screen. The main screen represents the basic information from which the Encore 2100 branches to provide more details as the user accesses other windows. 21 3.3. Systems Analysis The design conditions for this analysis are selected from CARRIER CODES and they are shown in Table 3.1. Refrigerant R-12 is used as the working fluid in the direct expansion cooling coils and the system operates on an ideal vapor-compression refrigeration cycle. And the refrigerant enters the compressor as a saturated vapor and the vaporization pressure is 20 psi. 3.3.1 Desiccant Cooling System In designing the desiccant cooling system, data gathered from the building were used as input. These data are shown in Table 3.2. In order to get the best design using desiccant units, some ofthe parameters which affect the performance of desiccant dehumidification were considered in the analysis and these parameters are [2]: 1. Process moisture 2. Process air temperature 3. Reactivation air temperature 4. Reactivation air moisture 5. Amount ofdesiccant presented to the reactivation and process air streams Since the model of desiccant wheel is chosen according to the amount of air entering to it [14], DC080 desiccant model is appropriate for this case, in which the process airflow is 7830 cfm. 22 Table 3.1 Design conditions- Main Sales Area 1. Summer a) Outdoor Design- Per State Energy Codes b) Indoor Design- 74 0 F ,D.B.- 50 0 F Dewpoint 2. Winter a) Outdoor Design- Per State Energy Codes b) Indoor Design-- 70 0 F .D.B. 3. People loads: 120 sq.ft./person of Sales Area (sensible and latent loads per ASHRAE) 4. Ventilation a) 15.0 C.F.M. Per Person Minimum b) Exact quantity of Ventilation Air shall be at least 10% greater than the total of Exhaust Air effecting the Sales Area system 5. General a) Conventional System- 0.75 to 0.85 C.F.M. per Square Foot of Floor Area b) Desiccant Dehumidification System- 0.65 to 0.75 C.F.M per Square Foot of Floor Area c) Miscellaneous heat producing equipment, within the design area, shall be considered in load calculation, unless directed not to do so. d) Refrigeration case credits 1) Case credits shall be used to in load calculations, unless directed not to do so 2) Credits shall be 50% of total case load for open faced multi-deck meat, multi deck dairy, produce, and frozen food cases. Of this 50%, 85% shall be sensible credits and 15% shall be latent credits 23 Table 3.2 Design Airflow Schedule/ Airflow Configuration Inlet Conditions Process Air CFM FDB FWB GRILB FDP RH Outside/Ambient 1400 92 72.4 88 63.6 39 Inside/Bldg Return 6430 75 64.1 72 58.1 55.6 Process Mixture 7830 78 65.7 74.9 59.2 52.4 Regeneration Air Outside/Ambient 7830 92 72.4 88 63.6 39 Inside/Bldg Exhaust 0 75 64.1 72 58.1 55.6 Regeneration Mixture 7830 92 72.4 88 63.6 39 Supply Air Conditions After Heat Exchanger 7830 82 60 42 44 27 24 The data were Implemented to the' Desiccant Unit Selection' software, and the process airflow leaves the desiccant wheel hot and dry, therefore an additional cooling will be needed to meet the sensible and latent load requirements. The heat wheel provides part of the sensible cooling. However, it is not sufficient to meet the sensible load requirement of the conditioned space. Therefore, indirect evaporative cooling coil can be used in this case. Consequently, in order to meet the load requirements of the conditioned space, the system must be composed of: desiccant rotor, heat exchanger rotor, boiler, regeneration heating coil, indirect evaporative cooler, process air blower, and regeneration air blower. Fig. 3.2 shows the desiccant unit selected and Table 3.3 shows the system data values for this design. 82 F m- J 7830 CfM 75 F U 116 g/lb - g- -fm- > w ~ OJ OJ C d .£ U x W OJ OJ .£ ::3 133 F 31 gM/UO 1t '25F 190 F U 104 g/Lb :r: -- +-' C d U U 1I1 OJ t=l C) co C) I U t=l OJ OJ .£ ::3 7830 CfM 78 F -« ~75 g/lb --- 0 _-.1---" f (i) 6430 erM 75 F 72 9M/llo 1400 efN 92 F 8.8~__ ---l£'.-J 43 gM/lb 1 j o --l 75 F 52 9M/1b _----@]--0~_-J- Fig. 3.2 Desiccant Cooling Unit, OH 25 ~ CO cff"l -« ~ efM erN 8.8~ ~ Table 3.3 System Data Values for the Desiccant System Component Dry Bulb Temp. Specific Humidity Airflow Process Air FDB GR/LB eFM Ventilation Air 92 88 1400 Vent-Return (Mix.) 78 75 7830 Desiccant Wheel 133 31 7830 Wheel Exchanger 82 43 7830 Zone Air 75 52 7830 Regeneration Air Outside/ Ambient 92 88 7830 Evap. Cooling Coil 75 116 7830 Wheel Exchanger 125 104 7830 Reg. Heating Coil 190 104 7830 26 3.3.2. Coefficient of performance for the desiccant cooling system The term coefficient of performance (COP) has been devised to measure the effectiveness of refrigerating machines and is usually defined as the ratio of the refrigeration produced to the network supplied. Usually the COP is used to analyze and compare cooling systems for their thermal performance. This definition of COP is applied mainly for conventional refrigerating machines. However, a definition provided by C. M. Shen and W. M. Worek [13] was used to define the COP for the all-cooling system including the system using desiccant, conventional or mixed systems. COP for the desiccant system is defined as [13], COP = cooling capacity = HaUl - H in energy input WI + W in2 + Q3 + w 4 (3.1) (Btu/hr) (3.2) mal = mixed air flow rate before the desiccant wheel hi = enthalpy ofmixed air (Btu/hr) (3.3) m az = supplied air flow rate h z = enthalpy ofsupplied air WI , w 4 = power input to run the supply and the regeneration fan Win 2 = power input to run the compressor ofthe evaporative cooler coil Q3 = power required to run the heating coil The energy required to run the heating coil equals to the energy required to heat the air 27 sensibly and is calculated from (3.4) Air-handling components m the systems such as fans, ducts, and so forth, are selected on the basis of volume flow rather than mass flow of air [3]. Therefore, ifthe air volume flow rate is to be determined, it is necessary to specify the point in the system where flow volume rate is to be determined and find the specific volume ofthe air at that point With the mass flow rate, m a, known and the specific volume of the air at the point, we may calculate the volume flow rate in cubic feet per minute from ,1. m aV cJm=-- 60 (3.5) For uniformity in the manufacturing industry air-handling equipment is normally rated on the basis of "standard air" which has been defined as dry air at 70° F and 29.92 in. Hg barometric pressure, which gives a density of 0.075 Ibmlft 3 [3]. Using the density, equal to l/VI in Eq. (3.5) and ifwe substitute the value of m a , we have (3.6) The power input to the compressor is calculated from: (3.7) m r =mass flow rate ofthe refrigerant (lb/hr) h 3 , h 4 =enthalpy ofthe refrigerant at conditions 3 and 4 (Btu/Ibm) Based on Eqs. (3.1) to (3.7), the value ofCOP for the desiccant system is calculated. Spreadsheets were used to find the value ofCOP as shown in Table 3.4 and also to graph the energy required to run the desiccant system as shown in Fig. 3.3. This energy is 28 composed mainly of thermal energy required to run the heating coils and electrical energy to run the compressors and fans. Graphing the energy required to run the systems in this or in the following cases gives an idea about the operating costs to run the systems. However, detailed cost analysis is not included in this study. 29 Table 3.4 COP for Desiccant, OH Enerav Reauired to Run the C Total Coil Refrigerant Inlet Compressor Inlet Evaporative Condensation Outlet Compressor Working Load (ton) Flow Rate (Lb/hr\ Enthalpy (Btu/lb) Enthalpy (Btu/lb) Pressure (psi) Enthalpy (Btu/lb) Input (Btu/hr) Evaporative Cooler 12 2300 76.4 13.8 39 81 10580 (,;J o Air Flow Specific Volume Enthalpy Of Air Energy Required Rate (cfm) (ft3/ lb ) (Btu/lb) (Btu/hr) Mixed Air 7830 13.75 30.5 1042102 Supplied Air 7830 13.75 26.5 854182 Desiccant Unit Evaporative Cooler 12 tons Heating Coil 548000 Btu/hr Supply Fan 10 Hp Return Fan 10 Hp 50890 The Coefficient Of Performance Mixed Air Supply Air Heating Coils Compressors Fans COP (Btu/hr) (Btu/hr) (Btu/hr) (Btu/hr) (Btu/hr) 1042102 854182 548000 10580 50890 0.308 E R"d R hCnerg~eqUire un t e ompressors (fe/lb) 600000 I --, 548000 500000 400000 ... .r;, '3 300000 -::=: m 200000 100000 0+1---- Heating Coils 10580 Compressors 50890 Fans Fig.3.3 Energy required to run the Desiccant System, OH ,-----------------------------------------, 0-+----- 3.3.3. Conventional Cooling System Most people are familiar with the principle of condensation. When air is chilled below its dewpoint temperature, moisture condenses on the nearest surface. The air is dehumidified by the process of cooling and condensation. The amount of moisture removed depends on how cold the air can be chilled- the lower the temperature, the drier the air. The moisture removed from the air during any dehumidification process may be determined directly from the difference in specific humidities for the actual entering and leaving states, while the latent heat removal associated with this moisture removal can be calculated from the following equation [4]: q/ =O.68(cfmX~W) (3.8) q, =latent heat removal (Btulhr) ~W = moisture removal (gr/lb) This is the operating principle behind most commercial and residential aIr conditioning systems. A refrigeration system cools air, drains away some of its moisture as condensate and sends the cooler, drier air back to the space. The system basically pumps the heat from the dehumidified air to a different airstream in another location, using the refrigerant gas to carry the heat. The actual hardware that accomplishes cooling dehumidification is exceptionally diverse. Literally thousands of different combinations of compressors are in use throughout the world. But there are three basic equipment configurations of interest to designers ofhumidity control systems [2], which include: • Direct expansion cooling • Chilled expansion cooling 32 O.68(cfmX~W) ~ • Dehumidification- reheat systems For the analysis, 'HAP' software was used for designing this system. Weather data, the space conditions and the properties of the building were chosen from ASHRAE based on the specifications and the sheets of the location and the properties of the building. Tables 3.5 and 3.6 show the design weather parameters and the space input data for the main sales area that were used during the design. The cooling of buildings is actually made up of two processes: sensible cooling, which is lowering air temperature, and latent cooling which is removing water vapor from the air. Cooling coils often have low latent capacities, usually ranging from 20% to 30%. This high coil sensible heat ratio can create problems when the SHR of the load falls below 70 %, since the coil will no longer have enough latent capacity to meet the latent load [14]. These cooling coils cool the air to levels between 43 and 45 0 F . Below that point, frost begins to form on parts of the coil, spreading slowly through coil as the airflow becomes restricted. The frost insulates the refrigerant from the air passing through the coil, which reduces heat transfer and the frost physically clogs the coil, reducing the airflow. Eventually the frost blocks the airflow all together and dehumidification ceases. So in this case, reheating the air must be performed in order to dehumidify it again to get the required humidity ofthe air concerened. In order to overcome this frosting problem and to meet the load requirements of the conditioned space, a dual air system was used. Air is supplied to the two air streams at different conditions (usually one hot, the other cold) and mixed by proportioning dampers upstream in a plenum. The entire air quantity for absorbing the load is conditioned centrally and distributed by the main fan. Mixing may be performed at the 33 Table 305 Design Weather Parameters for Youngstown, OH Design Weather Parameters & MSHGs Projed N"",e: k"",llI projedmlll Prep.red by AETOS ()esign Parameters City Name Youngstown Location Ohio Latitude 41.3 Deo Longitude 80.7 Deg Elevation 1184.0 n Summer Desion Dry-Bulb 88.0 OF Summer Coincident Wet-Bulb 72.4 OF Summer Dally Ranoe 20.6 OF Winter Design Dry-Bulb ·1.0 'F Winter Desion Wet-Bulb -2.5 'F Atmospheric Ciearness Number 1.00 Averaoe Ground Reflectance 0.20 Soil Conductivity 0.800 BTU/hrlfllf Local Time Zone (GMT +1- N hours) 5.0 hours Consider Daylight Savings Time Yes Daylioht Savinos Beoins AprH, 1 Daylioht Savings Ends October, 31 Simulation Weather Data 00000000000000000000 (00000) Current Data is User Modified Design Gig Months Januaryto December Design Day Maximum Solar Heat Gains (The MSHG values are expressed in BTU/hrll!") IMnnth N NNF NF FNF F FSF SF """" <: 1.I~nll~rv1R Q 1R Q 1R Q 7?Q UQ 1 1QR 1 71R7~1R7~4n I~ghrllorv?1 ?1~nn 17? R1~n?11 7 ?A7 Q ?An 7A11 I"",h ?R ?R Qn Q Inn?1~?~7?~n7 ?1Q 1 ?1 n 1 IAnril 114 flflS un 7 1BB~7n1n4~ ?M17~'~Qn 1M"" 1fl~~7~1fl1 fl 7n4 fl 717 7nQ <; 1771 11R 11 B 7 llolng A7<; 1nQ 1 171 n 7nfl1 7U 'nn 1 1fl1~ 171 <; 1n1 7 LIIIIv~7QQ n 1n1 71Q~'1 <; 'nA n 171 11A 1 11<;<; IAllnllct 1<; 1 flR 1 1~~1 IR??1~?1~IQR 1nQ 1~AA 7~7~A~nR 1~44 7nA~774 n nA1 717 Q ?n~ I()rfnhgr 'A 1 'A 1A~1 l1Q 7 17Q Q 77n n '1Q~71~<; )144 1Q 1 191 IQ 7n 1i?7 ?nn~?~?n?A~1 7AR R 1fl Q 1fl Q 1R~<;7 Q l1n 1 1R7 4 nA 3 74fl 4 2513 IMnnth <:<:\AI <:\AI W<:UII W WNW NW NNW I-IOR Mult. bnllo", 7<;11 73B5 ?n· UR nR 1R 1R Q17~~1 nn IF~hrllorv/4<; 7 74<; Q 711 4 1A7 fl 17R 1 4<; 7 734 171 1 nn IMorrh 71 Q1 71fl Q 717 '1<; R 1Rfl7 Ql~7R 4 '174 1 nn IAn,;1 171 ?nA 7 ?????~I Q1 nl~n~7n~?A7 7 1 nn 'Mav l1fl~ITT 7nT nnn~7nl~1fl7 Q 107 ?,,? Inn .IIIn~17n Q lfl41 1QQ 1 71fl1 7n1 171 117 7fl<; 1 nn .IIIIv l1A n 171 7n17 71<; " 1Q7 1R1 R 1nn 7<;Q R 1 nn All nilct '''R 1Q7 R ??1~1 RA 1~?nQ Q ?4~~1 nn 717 77S 0 77A fl 1QQ1~" Qn ?Q ?nQ 1 nn O,tnh~r71Q Q 74n 7 71 R 4 1R1 A 11<; 4R A 741fl~100 ?A7 7 717 R 1QQ 1 U7 n 77 1Q lQ 171 Q 1 nn ?A7 1 ??" Q IR" 17" ~Qn In 1R Q 1n4 R 1 nn Mutt.· User-defined solar muliplier fador. 34 O5I31AlO 10:18AM 1.I~nll~rv7~17~4 I~ghrllorv~n1~n ?1~?~7?~n ~n4~ 17~'~Q ~~7~ ~ ~71Q~ ~~?1~?1~~A 7~7~~n~4~ ?n~ A~1~71~ ~?~??A~ ~ 17~~ IF~hrllorv ~ ??~l~n~~ ~nn~~ .IIIn~ ?1~~?~~ 1~" O,tnh~r1fl~ ~Q Table 3.6 Space Input Data for the Giant Eagle Store Lc~s!.tp~a~ce~ln!.l:!p~u~tD~at~a'__~ Project Name Giani Eagle project Prepared by. AETOS Main Sales Area 'I. Gene. 31 DetQils: Floor Area 13601.0 fl' Avg Ceiling Height 14.000 fl Building Weight 10.000 Ibltt' 2.lnte.mtls: 2:1. Over head lighting: Fixture Type Recessed (Unvented) Waflage 2.50 Witt' Ballast Multiplier __1.00 SChedule I Ightino_overhead 2.2. Tau lighting: WtJltege -:1955.0 WtJlts SChedule Lightino_Task 2.). ElectJ ical E'luipment: Waflege 0.00 Witt' SChedule Hono 2.4. Peol>le: OCcupancy .J90.OO fl'lperson Activity Level User defined Sensible 250.0 BTUlhrlperson Latent 200.0 BTUmrlperson Schedule Occupenta 2.5. Miscellaneous LOAd", Sensible ,114020 BTUIhr Schedule Cabinets Lalent 0 BTllA'Y Schedule Cabinets ).1. Constluction TYI>es fOI Exposure S Wall Type DefaukWaiAsBembly 4.1. Constluction TYI>es fOI Exposure H Roof Type Assembly Roof 5. Illfittl mon: Design Cooling 0.0 CFM Design Healing 0.0 CFM Energy Analysis 0.0 CFM Infilratlon occurs only when the fan is off. 6. Floors: Type~SlabFloor On Grade Floor Area 1)601.0 fl' Total Floor U-Value 1.200 BTUIhrItt'1F Exposed Perimeter 113.0 fl Edge Insulalion R-Value 1.0 hr-fl'-FiBTU 1. PA.tition,,: (No part.Ion data) 35 ~s!.tp~a~ce~ln!.l:!p~u~tD~at~a'__~ ~Slab apparatus with only a single duct extending to the zone. To maintain conditions at all the times, the cooling conditioner and duct must be sized for the maximum cooling load that exists when no heating is required from the other duct, and vice versa. This means that for any system where there is a wide variation in the cooling and heating requirements among the spaces served, each conditioner and duct will have to be sized to carry possibly 75% or more of the maximum load. At peak summer load, both air streams can carry the cooling effect, provided the equipment is arranged to furnish it [4]. Fig. 3.4 shows the conventional unit selected and conditions of air ® 4453 cf.., 48 F 40.8 9..,/10 [i~~";,J .-_. _..-.-@l--l- -- .. -1..:12)--- -----.-...~--_....J Fig. 3.4 Conventional Cooling Unit 36 [i~~";,J ~--_.. In order to meet the load requirements of the conditioned space, the system as shown in Fig. 3.4 must be composed of pre-cooling coil, pre-heating coil, a supply air-blower, a return air-blower and a dual duct system that contains heating and cooling coils and a mixing box. System data values that are shown in Fig. 3.4 were taken from Table 3.7 and the psychrometric analysis for the conventional cooling unit is shown in Fig.3.5. Table 3.7 System Psychrometric for the Conventional Unit System Psychrometries for Packaged Rooftop AHU Projed N4me: GIant Eoglc project Prepared by; AETOS JulY DESIGII COOLUIG DAY, 1100 TABLE l' SYSTEM DATA Dry..sulb Specific SenllllJIeL~ent Temp lIumldlty AIrflowlle~He~ omDF: 9191 30533 - TABLE 1 ZOIlEDATA: Zone Tcmlinal Zonc Sensible Zone Zono Zono Hellting Heating Load T·_ Cond Temp AltfIow Coil Unit 70nollam" ,nTUIh,· Mode 'BTII"''' ,of, ICFMI IBTUIh'I 'BTUlhrt ZOI1e 1 1AA.1!4 Delldbsnd 0 73.6 9191 0 0 37 L~ent lle~He~ omD ..._._~---------_.--' Fig. 4.6 Alternative System Design, OR In order to meet the particular load requirements, the system must be composed of pre-cooling coil, desiccant rotor, heat exchanger rotor, boiler, regeneration heating coil, process air blower and regeneration air blower from the desiccant side. Also, it is composed of cooling coil, heating coil and a return air blower. The COP for this system is defined as [13], Haul -H;n COP =------=-="------"'---- mal' m a2 =ventilation and return air flow rate 56 (4.2) (4.3) l~O Table 4.4 System Data Values for the Design System, OH Component Dry Bulb Temp. Specific Humidity Airflow Process Air FDB GRILB CFM Ventilation Air 92 88 1400 Pre-Cooling Coil 68 83 1400 Desiccant Wheel 150 24 1400 Wheel Exchanger 75 29 1400 Mixing Air 75 64 7830 Cooling Coil 50 43 7830 Heating Coil 82 43 7830 Zone Air 75 52 7830 Regeneration Air Outside/ Ambient 92 88 1400 Wheel Exchanger 143 78 1400 Reg. Heating Coil 190 78 1400 57 h, ,h 2 = enthalpy ofventilation and return air H·::.:m )hl, in a_ rna) = supply air flow rate h) = enthalpy ofsupply air (4.4) Win' ' Win 4 = power input to run the compressors ofthe pre-cooling and cooling coils Q2' Qs = energy required to run the regeneration-heating and heating coils W 3 = power input to run the fans Based on Eqs. (3.4) to (3.8) and Eqs. (4.2) to (4.4), the COP for this system was found to be 0.302. The calculation was done on Excel spreadsheet which is shown in table 4.5. Fig. 4.7. shows the energy required to run the system. 4.4.2. Desiccant unit, FL For the desiccant unit, the same DC020 model was chosen based on the same air flow rate. New data were implemented using the same software used above. Properties of the building and space data are the same as in Ohio, but weather data are different. In order to meet the load requirements of the conditioned space, the system must be composed ofthe same elements as those in Ohio. Fig. 4.8. shows the system selected and table 4.6. shows the system data values for this design, 58 Table 4.5 COP for the New Design System, OH Eneray Reauired to Run the C Total Coil Refrigerant Inlet Compressor Inlet Evaporative Condensation Outlet Compressor Working Load (ton) Flow Rate (Lb/hr\ Enthalpy (Btu/lb) Enthalpy (Btu/lb) Pressure (psi) Enthalpy (Btu/lb) Input (Btu/hr) Pre-Cool Coil 3.5 704 76.4 18 54.8 82 3942 Cooling Coil 27 6270 76.4 24.8 89.4 87 66462 lJ\ \Q Air Flow Specific Volume Enthalpy Of Air Energy Required Rate (cfm) (ft3 /Ib ) (Btu/lb) (Btu/hr) Ventilation Air 1400 14.15 35.9 213117 Returned Air 6430 13.7 29.3 825105 Supplied Air 7830 13.8 26.4 898748 (The Coefficient Of Performance Mixed Design Unit Pre-Cool Coil 3.5 tons Coolina Coil 27 tons Reg-Heating Coil 70000 Btu/hr HeatinQ Coil 270605 Btu/hr Fans 20 Hp Ventilation Returned Air Supply Air Heating Coils Compressors Fans COP Air (Btu/hr) (Btu/hr) (Btu/hr) (Btu/hr) (Btu/hr) lBtu/hr) 213117 8251051 898748 340605 70404 50890 0.302 E R 'dR hCnergoy eqUire un t e ompressors (Btullb) (Btullb) (Btullb) (Btulhr) (Btullb) (Btulhr) Heating (Btulhr) (Btulhr) (Btulhr) 50890 70404 340605 o +1-------1 100000 50000 150000 250000 350000 -j~ I 300000 ... ~ :; 200000 - a:I ~ Heating Coils Compressors Fans Fig4.7 Energy required to run the Alternative System, OR 400000 I ~ 0-1----- I 16\ .- 12 gf"l/tb -@-J 1<0' C'" Fig.4.8 Alternative System Design, FL Based on Eqs. (3.4) to (3.8) and Eqs.(4.2) to (4.4), the COP for this system is 0.3573. The calculation was done on Excel spreadsheet which is shown in Table 4.7. Fig. 4.9 shows the energy required to run the entire system. 61 Table 4.6 System Data Values for the Design System, FL Component Dry Bulb Temp. Specific Humidity Airflow -- Process Air FDB GRILB CFM Ventilation Air 92 116 1400 Pre-Cooling Coil 69 92 1400 -- Desiccant Wheel 161 32 1400 Wheel Exchanger 77 37 1400 - Mixing Air 75 66 7830 Cooling Coil 50 43 7830 Heating Coil 82 43 7830 Zone Air 75 52 7830 Regeneration Air Outside/ Ambient 92 116 1400 Wheel Exchanger 153 87 1400 Reg. Heating Coil 190 87 1400 62 Table 4.7 COP for the New Design System, FL Enerav Reauired to Run the C - - - - -- Total Coil Refrigerant Inlet Compressor Inlet Evaporative Condensation Outlet Compressor Working Load (ton) Flow Rate (Lb/hr) Enthalpy (Btu/lb) Enthalpy (Btu/lb) Pressure (psi) Enthalpv (Btu/lb) Input (Btulhr) Pre-Cool Coil 5 1029 76.4 18 54.8 82 5762 Cooling Coil 28 6504 76.4 25 90.6 88 75446 0 W Air Flow Specific Volume Enthalpy Of Air Energy Required Rate (cfm) (fe/lb) (Btu/lb) (Btu/hr) Ventilation Air 1400 14.25 40.3 237558 Returned Air 6430 13.7 29.3 825105 Supplied Air 7830 13.8 26.4 898748 MixedDesi~n unit (FL) Pre-Cool Coil 5 tons CoolinQ Coil 28 tons Req-Heating Coil 56000 Btu/hr Heating Coil 270605 Btulhr Fans 20 Hp 50890 The Coefficient Of Performance Ventilation Returned Air Supply Air Heating Coils Compressors Fans COP Air (Btu/hr) (Btu/hr) (Btu/hr) (Btulhr) (Btulhr) (Btulhr) 237558 8251051 898748 326605 81209 50890 0.3573 E R 'dR hCnergy eqUire un t e ampressors Desi~ 350000 I I 326605 ~ ... .J::. ~ - In 300000 250000 200000 150000 100000 50000 0+1---- Heating Coils 81209 Compressors 50890 Fans Fig. 4.9 Energy Required to run the Alternative System, FL 350000,.---------------------------------------------, 0+---- 4.5. Comparison between the systems 4.5.1. The original desiccant system and the alternative design in Ohio The value of COP using the new design system in OH is 0.302 and that value USlllg the original desiccant system is 0.308. Therefore, the system will perform better using the desiccant system alone in these kinds ofstores in OH. The energy required to run the two systems is shown in Fig. 4.10 for easy comparison. This does not necessarily mean that the system using larger amount of energy costs more in the operation because the thermal energy is much less expensive than the electrical energy. 4.5.2. The original desiccant system and the alternative design in Florida The value of COP using the new design is 0.3575 and is more than that using the desiccant unit alone which is 0.3508. Therefore, the new design system performs better than using the desiccant units in these kind ofstores in humid areas. Fig.4.11. shows the energy required to run the two systems 65 600000 I I 548000 0' 0' 500000 j 400000 ... ..c: "3 300000 m 200000 100000 0+1-- Heating Coils 70404 10580 Compressors 50890 50890 Fans Iii Desiccant, OH \_Design Improvement Fig.4.10 Energy Required to run Desiccant, OH and its Alternative Design 600000 -r------------------------ --, 400000 ... ..c: "3 100000 0+--- Iii Design Improvement 600000 I I 518000 0 -.J 500000 400000 ... .c ~300000 -In 200000 100000 0+1-- Heating Coils 81209 7841 Compressors 50890 50890 Fans IIDesiccant, FL : • Design Improvement i Fig.4.11 Energy Required to run Desiccant, FL and its Alternative Design ,--------------------------------------, ~ 0+--- CHAPTER V RESULTS AND CONCLUSIONS 5.1. Results The results of this analysis are summarized as follows: 1. The coefficients of performance for the system in OH, which is located in an area of 88 gr/lb specific humidity, is 0.308 using the desiccant unit and 0.231 using the conventional cooling unit, respectively. 2, The coefficient of performance of the desiccant unit, which is located in an area of 116 gr/lb in FL, is 0.351. 3. For the combined system of desiccant and conventional units, the coefficients of performance of the system in OH and FL are 0.302 and 0.357 respectively. For a better illustration, the COPs are plotted in Fig.5.l. 68 J o -t-!--- 0.2 0.1 0.15 I 04 1 0.3573 I 0.3508 I i O.~I 03 1 II iii I 0 2307 I I I I . I I 0.05 0.25 0\ '" Desiccant Unit, OH Conventional Unit, OH Alternative Design, OH Desiccant Unit, FL Alternative Design, FL Fig.5.l Coefficient of Perfonnance 0+--- 1 0.3573 0.35 03j 0.3083 0.302 I 0.2307 5.2. Conclusions Evaluations made from the study draw general conclusions on the performance of the systems: 1. Desiccant units perform better than conventional units in an environment of large humidity such as supermarkets. 2. The desiccant units in humid areas perform better than those in less humid areas. 3. Pre-cooling the make-up air and dehumidifying it with the desiccant before the air blended with return air from the zone must be considered when: a) The latent load of the make-up air is much larger than that of the internal moisture load ofthe zone. b) The system requires a large proportion ofmake-up air. 4. Desiccants are especially efficient when drying air to create low relative humidities, and cooling-based dehumidification is very efficient when drying air to saturated air conditions. If the air passing through the desiccant process is still close to saturation point at a lower temperature, cooling-based dehumidification would be a good choice. But if the desired end result is the air at a condition far from saturation, which means a low relative humidities, a desiccant unit would be ideaL 5, Desiccant cooling units perform better than conventional cooling units when the latent load is large in comparison to the sensible load. The cooling coils often have low latent capacities, usually ranging 20% to 30%, which means higher sensible heat ratio (SHR). Therefore, this condition can create problems to the 70 cooling coil when the SHR of the load falls below 70%, since the coil will no longer have enough latent capacity to meet the latent load. 71 BIBLOGRAPHY Books 1. ASHRAE Handbook of Fundamentals. The American Society of Heating, Refrigeration and Air Conditioning Engineers, Inc. New York: ASHRAE, 1989. 2. Cargocaire. The Dehumidification Handbook. Second Edition. MA: Munters Cargocaire, 1990. 3. Clifford, George E. Heating, Ventilation and Air Conditioning. Virginia, Reston Publishing Company, Inc. 1948. 4. Carrier, Willis H., Cherne, Realto E., Grant, Walter A and Roberts, William H. Modem Air Conditioning, Heating, and Ventilation. Third Edition. New York: W. H. Carrier, R.E. Cherne and W.A Grant, 1959. 5. Grimm, Nils R. and Rosaler, Robert C. Handbook of HVAC Design. New York: McGraw-Hill, Inc. 1990. 6. Jones, W.P. Air conditioning Applications and Design. England, Butler and Tanner Limited, Frome, 1980. 7. Porges, F. HVAC Engineer's Handbook. Tenth Edition. England, F. Porges, 1995. 8. Cengel, Yunus A and Boles, Michael A Thermodynamics An Engineering Approach. Second Edition. New York: McGraw-Hill, Inc. 1994. 9. Cengel, Yunus A. Heat Transfer A Practical Approach. New York: McGraw-Hill, Inc. 1998. 10. James, Ronald W. Desiccants and Humectants. New Jersey: Noyes Data Corporation 1973. Articles 11. ORNL R&D on Desiccant Technologies, Desiccant Research & Development http://www.orni.gov/ORNL/BTC/desiccant.html 12. Belding, William A, Delmas, Marc P.F. and Holeman, William D., Desiccant Aging and its Effect on Desiccant Cooling System Performance. England, Elsevier Science Ltd,1996. 13. Shen, eM. and Worek, W.M. The Second Law Analysis of a Recirculation Cycle Desiccant Cooling System. England, Elsevier Science Ltd, 1996. 72 14. Fresh Air SolutIOns, DS-2P Catalogue, 1998 http://wviw.freshairsolutions.com 15. RJ. Karg Associates, Natural Gas Marketing Articles htJpj/wwvi.karg.comigasmarket.htm 73 APPENDIX Fig. A.I Picture of a Desiccant Cooling System 74 Fig. A.2 Process and Reactivation airflow temperature and humidity changes 300 Reactivation heater Air mOIsture Air temperature 122 ---,- Process air i I I -------+1---'------- 110 °1 257 I I I 1 i Honeycombel' dehumidifier i I I , --r Ilooi 90 120 I I ! I ! I --L.... Legend -KEML':£ Cr/lb --- 10 200 150 250 c: o 'i 100 +----o/f > B fl::S 50 IV '- !:I: < Air temperature Air mOisture JL5JU.2Si. 56 150 100 50 10 75 Table A.I Cooling Design Temperature Profiles Cooling Design Temperature Profiles Project Name: Giant Eagle project Prepared by: AETOS locillion: Youngllto.....n. Ohio ( Dry and Wet Bulb temperatures are expressed in·F ) Hr Januarv Februarv Marcil ADril MiIII June DB we DB we DB we DB WB DB WB IlR WR 0000 305 300 34.5 34.0 461 45.6 57.3 56.8 663 624 733 657 0100 295 290 33.5 330 451 446 56.1 556 651 61.9 721 65.3 0200 7A.4 7A 0 17.4 11.9 44.0 43.6 55.1 54.6 64.1 61.6 711 650 0300 276 271 316 311 432 427 54.0 53.6 63.0 61.2 700 646 0400 270 265 31.0 30.5 42.6 42.1 532 52.7 62.2 61.0 69.2 64.4 0500 26.8 26.3 308 301 424 41.9 52.6 52.1 61.6 60.7 68.6 642 0600 272 26.7 31.2 307 42.8 42.3 52.4 51.9 614 607 68.4 641 0700 28.2 27.7 32.2 31.7 438 43. 528 52.3 61.8 60.8 68.8 64.2 0800 301 29.6 34.1 33.6 45.7 45.2 53.8 533 628 61.2 698 646 0!l00 17.A 17. 16.A 16. 4A.4 47.!l 55.7 55. 64.7 61.8 71.7 65.2 1000 35.9 354 39.9 39.4 51.5 51.0 584 57.2 67.4 62.7 74.4 66.0 1100 394 38.9 434 42.9 550 54.5 61.5 58.3 70.5 63.8 77.5 67.0 1200 42.7 42.1 46.7 46.2 58.3 56.6 650 59.6 74.0 64.9 810 68.0 1300 45.1 43.3 49.1 48.4 607 57.6 683 608 77.3 65.9 843 69.0 1400 46.8 44.1 50.8 4!l.1 674 5A. 707 616 7!l.7 66.7 A6.? INA 1500 47.4 444 51.4 49.4 630 58.4 724 62.2 81.4 67. 88.4 702 1600 46.8 44.1 50.8 49.1 624 58.2 730 624 82.0 67.4 890 704 1700 45. 43.4 49.3 48.5 609 57.6 72.4 62.2 81.4 67.2 88.4 70.2 1800 431 423 47.1 46.6 587 56.8 70.9 61.7 79.9 66.8 86.9 69.8 1900 40.4 39.9 444 41.!l 56 IJ 55.5 6A.7 609 777 66.1 A4.7 6!l.7 2000 37.7 37.2 41.7 41.2 53.3 52.8 66.0 60.0 75.0 65.2 820 68.4 2100 355 35.0 39.5 39.0 51.1 50.6 63. 59.0 72.3 64.3 79.3 67.5 2200 33.4 329 37.4 36.9 490 485 61.1 58.2 70.1 63.6 77.1 669 7100 31.7 31.2 35.7 35.7 47.3 46.8 590 57.4 680 629 75.0 662 Hr Jltv Aucust SeDtember October NOII'ember December DB we DB we DB we DB we IlR WR IlR WR 0000 76.1 67.9 76.3 67.9 703 64.6 60.3 59.0 485 48.0 36.5 360 0100 75.1 675 75.1 67.5 691 64.2 591 58.6 47.5 47.0 35.5 35.0 0200 74.1 67.2 74.1 672 681 63.8 58.1 57.6 46.4 46.0 34.4 33.9 0300 730 66.9 730 66.9 670 635 570 56.6 456 45.1 33.6 33.1 0400 72. 66.6 72.2 66.6 66.2 63.2 56.2 55.7 45.0 44.5 110 17.5 0500 71.6 664 71.6 66.4 656 63.0 55.6 55.1 44.8 443 32.8 323 0600 71.4 664 71.4 664 65.4 63.0 554 54.9 452 44.7 332 32.7 0700 71.8 665 71.8 665 658 63.1 558 553 46.2 45.7 34.2 337 0800 72.8 66.8 728 668 66.8 63.4 568 56.3 48.1 47.6 361 356 0900 74.7 67.4 74.7 674 687 64.0 587 58. 50.A 511.1 3AA 1A.1 1000 77.4 682 77.4 68.2 71.4 64.9 61.4 59.4 53.9 53.4 41.9 41.4 1100 80.5 6!l1 AO 5 6!l.1 74.5 65.9 64.5 60.5 57.4 554 45.4 44.9 1200 84.0 702 84.0 702 780 67.0 680 61.7 607 56.6 48.7 473 1300 87.3 71.1 87.3 71.1 813 68.0 713 628 631 576 511 48.4 1400 8U 71.A 89.7 71.A A3.7 6A7 737 61.7 64.8 58.2 52.8 4!l.1 1500 91.4 72.2 91.4 77.2 85.4 69.2 754 64.2 65.4 584 534 49.4 1600 92.0 72.4 92.0 72.4 A60 69.4 76./1 64.4 64.8 58.2 528 49.1 1700 91.4 722 91.4 722 854 69.2 754 64.2 63.3 576 51.3 485 1800 89.9 718 89.9 718 839 68.8 739 63.7 61.1 568 491 47.5 1!lIJO A7.7 71 A7.7 71./ AU 68.1 71.7 61 IJ 5A 4 55.8 46.4 459 2000 850 70.5 85.0 70.5 79.0 67.3 69.0 62.1 557 54.7 43.7 432 2100 82. 69.7 82.3 69.7 76 66.5 66. 61.1 53.5 53.0 41.5 41.0 2200 80.1 690 801 69.0 74.1 65.8 641 60.4 51.4 50.9 394 38.9 2300 780 684 780 68.4 720 65.1 620 59.6 497 49.3 37.7 37.2 76 Fig. A.3 Design temperature profile c=- Design Temperature Profile Project Name: Giant Eagle project Prepared by:AE~OS Loc.ltion: Youngstown, Ohio Design Temperature Profiles for July I "I Dry 8ulb W~8ulb , , , , , , , , , , --1--------·---.------------,--------- , , , , , , , , , , , , , , , , , , , , , , , , , , , , , , , , , , , , I I I I - - (" - - - - - - - - - - - - r - - - - - - • - - - - - r - - - - -: .• - - - - - r - - - - - - - - - - - - ..~- - - - - - - - .. , , , , " , " , " , " , " , " , " , " , " , " , I I I , I --r------------r-----~------r-----------,------------r---- I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I , " , " I I I I I I I I I --~----------~----------~------------~------------~--------- I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I I , , , , , , , , , , , , , , , --~------------~------------~--- , , , , , , , , , , , , , , , , , , .,---r-------.,--------,-------'"T'"----~.----- , , , , , , , , , , -1------ ----.---- 85 90 70 ~ 80 ~ .3 ro iii c. E Cll f- 75 o 5 10 15 20 Hour 77 AE~OS W~8ulb ~ --r------------r-----~------r- --~----------~----------~------------~------------~--------- --~------------~------------~--- .,---r-------.,--------,-------'"T'"----~.----- ~ ~ Table A,2 Air System Sizing Summary for the Conventional Unit '-- ....:.A..::.i=--rs;::'YL.s:::..:t:.::ce.:.:.m.:-S=..io=zing S'ummary for Packaged Rooftop AHU Projeel Name Giant Eagle projeel Prepared by: AETOS Ail SystemlllfOIIll,ltion System Name -'Pack.ged Rooftop AHU Equipment Oass PKG ROOF System Type DDCAV Sizing C....lculmioniidollllmion Zone and SI).ce Sizing Method: Zone CfM~Pe.kzone 8ensible lo.d Space CFM Coincident space lo.ds Celltl ....1Cooling Coil Sizing Om.... Total coi load 14..6 Tons sensible coli load 12.0 Tons Coil CfM at Jun 2200 4595 CFM Max possible CFM 9U1 CFM Design supply lemp. 41.0 "F fl'fTon 932.6 BTUAvIfl' 12.9 wmer flow @ 10.0 'F rise - gpm Cellt,at Heating Coil Sizing Oata Max coli load 122541 BTUAv ColI CfM at Des I-Ig US6 CFM Max possible CFM US6 CFM Water flow @ 20.0 "F drop - gpm PI ecool Coil Sizing 0,11,1 Total coi load -"22.9 Tons Sensible coil load 13.1 Tons Col CFM at Jul1600 3180 CFM Max possille CFM 3710 CFM Water flow @ 10.0 'F rise • gpm PI eheat Coil Sizing Om.... Max coi load --'96OC9 BTUn ColI CFM at Des I-Ig 9191 CfM Max possible CFM 9191 CFM Water fIow@ 20.0 'F drop • gpm Supply Fan Sizing 0,1t.l Aelual max CFM at Jul1700~9191CFM standerd CFM 1104 CFM Aelual max CFMIfl' 0.61 CFMIfl' Retulll Fan Sizing OmiI Aelual max CFM at Jul1700 .9191 CFM standard CFM 1104 CFM Aelual max CFMIfl' 0.68 CFMIfl' Outdool Velltiimioll Ail 0,1t,1 Design airflow CFM~3180CFM CfMIfl' 0.28 CFMIfl' 78 Number of Zones 1 Floor Area 13601.0 fl' Calculation Months ....J.n to Dec Sizing Data C.lculated Load occurs at -"Jun 2200 OA DB/lMl 17.1/66.9 "F Entering DB/V1113 18.2/60.1 "F Leaving DB IV1113 41.0 145.0 'F Coli ADP 40.4 Of Bypass faelor 0.200 Resutting RH 41 % Zone T-stat Check 1 of 1 OK Loed occurs at -"Des Htg BTUklrlfl' 9.0 Ent DB 1Lvg DB 18.2 I ".0 "F Load occurs at .....JuI1600 OA DB/lMl 92.0112.4 'F Entering DB/'MJ 92.0112.4 "F Leavi"lg DB IV'IfJ 50.0141.0 "F Bypass faelor 0.200 Load occurs at oPDes Iltg Ent. DB ILvgDB 6U/lS.0"F Fan motor BHP 12.00 BHP Fan motor 'tNV 8.95 'tNV Fan motor BHP 12.00 BHP Fan motor 'tNV 8.95 'tNV ~Pe.k eCool ~9191 ~3180