Abstract:
A multiple stage dehumidification and cooling system wherein a first stage direct expansion dehumidifier operating at its optimum dew point to an entering high humidity, high temperature air stream and effecting a first lowering of the temperature and humidity of the air stream, with the conditioned air stream being serially conveyed to a second stage chilled liquid dehumidifier operating at its optimum dew point to effect a second lowering of the temperature and humidity, and thereafter to a third stage reheat coil for providing an exiting air stream of desired temperature and humidity conditioning. The synergistic coupling provides significant power saving over the prior alternatives of desiccant and chilled liquid systems and the stages are individually modulated to reduce power consumptions as the load temperature and humidity set points are approached.

Description:
RELATED APPLICATION 
     This application claims the benefit under 35 USC 121 of United States Provisional Application No. 60/358,685 filed on Feb. 21, 2002 in the name of Thomas J. Backman and entitled “Hybrid Dehumidifier and Cooling System. 
    
    
     FIELD OF THE INVENTION 
     The present invention relates to apparatus for cooling apparatus and, in particular to a dehumidification and cooling system employing synergistic effects of serially coupled direct expansion (“DX”) and liquid chilling to achieve low air stream dew points with low power consumption under conditions of high moisture loads. 
     BACKGROUND OF THE INVENTION 
     The cooling systems commercial and retail facilities generally include a remotely located primary unit that is individually connected to various cooling loads or zones, such as air conditioning. Chilled liquid or direct expansion cooling systems are typically used. 
     Evolving standards and regulations are requiring increased outdoor air introduction into commercial and industrial buildings for improving interior air quality. Introducing such outdoor air: into areas having stringent humidity control requirements can greatly increase of dehumidification removal load requirements, particularly during periods of increased temperature and humidity. Humidity sensitive environments such as supermarkets, libraries, sports arenas, hotels, food storage, and process control areas can suffer severe adverse operational problems, from mold, mildew, and product and equipment damage if the cooling systems cannot handle the increased moisture. To adequately handle moisture removal in such situations, it has been widely accepted that an internal dew point temperature of 50° F. or less is required in these spaces, and that the supply air accordingly must be about 40° F. At such lowered temperature, traditional direct expansion dehumidifiers are prone to icing, and supplemental defrost systems are required. The additional costs associated with the defrost systems and the attendant operational problems have reduced the use of direct expansion dehumidification systems in these humidity dependent applications. 
     The lower dew points can be achieved without defrost cycles using chilled liquid systems, enabling operational dew points as low as about 34° F. Sophisticated controls systems, however, are required and the power consumption is greater than the direct expansion systems. Alternatively, desiccant dehumidifiers may be used to achieve these requisite dew point conditions, but only at high operational and maintenance costs. 
     BRIEF SUMMARY OF THE INVENTION 
     The present invention provides a multiple stage dehumidification and cooling system wherein a first stage direct expansion dehumidifier operating at its optimum dew point to an entering high humidity, high temperature air stream and effecting a first lowering of the temperature and humidity of the air stream, with the conditioned air stream being serially conveyed to a second stage chilled liquid dehumidifier operating at its optimum dew point and effecting a second lowering of the temperature and humidity, and thereafter to a third stage reheat coil for providing an exiting air stream of desired temperature and humidity conditioning. The synergistic coupling provides significant power saving over the prior alternatives of desiccant and chilled liquid systems. Further, the stages are individually modulated to reduce power consumptions as the load temperature and humidity set points are approached. 
    
    
     DESCRIPTION OF THE DRAWINGS 
     The above and other objects and advantages of the invention will become apparent upon reading the following written description taken in conjunction with the accompanying drawings in which: 
     FIG. 1 is a block diagram of the mechanical components for the multiple stage dehumidification system in accordance with a preferred embodiment of the invention; and 
     FIG. 2 is a block diagram of the dehumidification system including the control system. 
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
     Referring to the drawings for the purpose of illustrating a preferred embodiment of the invention and not for limiting same, FIG. 1 shows a multiple stage dehumidification and cooling system  10  including an air handler  12  for receiving an input air stream  14 , from interior and/or exterior sources and at variable temperature and humidity conditions, and delivering an output air stream  16  for routing to an environmental load  18  to establish and maintain predetermined temperature and humidity conditions thereat. The system may operate as freestanding units for cooling and dehumidification requirements, or as a secondary unit for handling extreme environmental conditions. 
     The air handler  12  comprises a housing  20  defining an internal fluid passage  22  having an inlet  24  fluidly coupled with the inlet stream  14  and an outlet  26  fluidly coupled with the output stream  16 . Serially disposed in the passage  22  downstream of the inlet  24  are a direct expansion heat transfer coil  30 , a chilled liquid heat transfer coil  32 , a reheat coil  34  and a delivery fan  36 . The housing is provided with a drain  37  for removing condensed moisture. 
     The inlet stream  14  may be furnished from the load  18  and/or outside air  15  through conventional valving  17 . The system as hereinafter described has particular application in situations wherein regulations or other consideration require substantial quantities of outside air that may further increase the system demands. 
     The direct expansion heat transfer coil  30  is disposed in a direct expansion thermal cooling loop  40  serially connected in the direction indicated by the arrows to a compressor  42 , a condenser  44  and an expansion control valve  46 . The components for the loop  40  are well known in the art and sized and selected in accordance with the requirements of the load. The loop  40  employs any suitable direct expansion refrigerant, for example R-22, or a refrigerant on the list hereinafter set forth. 
     The liquid heat transfer coil  32  is connected in a plural stage loop comprising a primary loop  50  and a secondary loop  52 . The coil  32  is disposed in the secondary loop  52  and serially connected in the direction indicated by the arrows to a liquid chiller  54  and a liquid pump  56 . The liquid chiller  54  is also connected in the primary loop  50  serially with a compressor  58  and a condenser  60 . The loops  50  and  52  employ any suitable liquid refrigerant, for example glycol, or accepted refrigerants on the list hereinafter set forth. 
     The reheat coil  34  is serially connected in a reheat loop  62  with any suitable source providing suitable heating capacity, such as the waste lines of the liquid compressor  54 . 
     The above coils  30  and  32  are heat transfer systems for lowering the air stream to a temperature below the dew point. The lowest mean temperature reached is called the air dew point temperature. During cooling, water vapor condenses on coil fins and the liquid routed to liquid drain. Exiting the coils, the air stream is cold and saturated (100% relative humidity) with water. Thereafter, heat is introduced to the air stream at the reheat coil to increase the air temperature with a resultant lowered humidity. 
     Direct expansion cooling coils are more energy efficient in dehumidifiers than chilled liquid cooling coils because the heat transfer is effected in a direct primary loop and does not require the secondary loop, including an additional heat exchanger and a direct expansion system which have incremental power consumptions. 
     When operating a dehumidifier, the system is designed to keep the coil temperature above 32° F. at all times to avoid forming on the cooling coil and blocking the airflow therethrough. This is achieved by operating the cooling coils at a gradient of about 10° F. below the exit temperature of the air. Additionally, a fluctuation safety factor of about 4° F. is incorporation such that that DX dehumidifiers typically have a minimum threshold air temperature of 46° F. establishing the minimum moisture level in the air stream exiting the expansion coil, hereinafter DX optimum operating dew point. 
     Most chilled liquid cooling coils are designed with about a 1° F. temperature gradient and a 1° F. fluctuation safety factor thereby establishing a minimum exit temperature from cooling coils of about 34° F. or CL optimum operating dew point. The actual operating optimum dew points may vary slightly based on manufacture and design, but each will have an accepted operational temperature that prevents freezing and icing experiences. It will thus be appreciated that while the chilled liquid coil is less efficient than a DX coil, it can create much lower humidity that DX coils. 
     The foregoing design and control of the present system synergistically takes advantage of the high efficiency of DX cooling in conjunction with the low dew point of chilled liquid cooling. As hereinafter described, the DX coil cools the air stream to 46° F. serially followed by the chilled liquid coil cooling the air stream further to a temperature of 34° F. Reheat as required delivers an efficiently dehumidified air supply to high demand spaces. The individual stages are controlled to provide progress power reductions for the overall system as the load approaches humidity and temperature set points. 
     The two cooling systems in serial air stream arrangement achieve cooling and dehumidification in a combination of low power consumption with low dehumidification dew point that cannot be achieved with either direct expansion systems or liquid secondary cooling systems operating alone. The system takes advantage of the high dehumidification capabilities of both systems. 
     Referring to FIG. 2, the control system  80  for the dehumidification system  10  comprises a process controller  82  that determines local conditions at the load  84  with a load temperature sensor  86  and a load humidity sensor  88 . The controller  82  is interfaced with a waste coil temperature sensor  90  downstream of the reheat coil  42 , a chilled coil temperature sensor  92  downstream of the chilled coils  44 , and an expansion coil sensor  94  downstream of the expansion coil  46 . The controller  80  is further interfaced with the variable speed fan  48  and the expansion valve  62  in the expansion loop  40 . 
     During an “unoccupied mode”, the system  10  is disabled through an appropriate command at the controller  80 . When the system is enabled, the prevailing conditions at the load  18  are determined by the temperature sensor  82  and the humidity sensor  84 . If either is outside the set points, the fan  48  is operated to draw the air stream through the air handler. Thereafter, if the temperature sensor  84  is low, the reheat coil  48  is energized to control the exit air temperature from the reheat coil at the target temperature. If the load temperature is within limits, the reheat coil remains disabled. If the humidity is above the set point at the load humidity sensor  84 , the chilled liquid system is enabled if the exit temperature sensed by sensor  90  is above the CL optimum dew point temperature. If the exit temperature is below the threshold value, the chilled liquid system remains disabled. After stabilization of the chilled liquid system, the controller  80  polls the expansion coil sensor  44 . If the exit temperature is above the DX optimum dew point temperature value, the expansion system is enabled. If below, the expansion system remains disabled. In this fully operating mode, the dehumidification is handled predominantly by the direct expansion stage. 
     Under operating conditions, as the load humidity approaches set point, the expansion valve  46  is progressive throttled to maintain the optimum dew point temperature value at the transient lower demand, which would otherwise result in an excursion therebelow and icing of the coils. Under further reductions, the expansion valve is further throttled until closed resulting in progressive power savings, and the controller disables the expansion system. Thereat, the residual humidity load is handled by the chilled liquid system. Upon further humidity decreases in the inlet air, the controller  82  decreases the speed of the fan to decrease flow rate through the handler  14  while maintaining the CL optimum dew point temperature value and progressively continuing until the set point humidity at the load is attained. For subsequent operational transient, the control sequences are reversely adopted. As a result, the phased dual mode humidification provides humidity control without reheat or evaporation modalities, allows for capital downsizing of single mode systems and importantly reduces operating costs in comparison with the now single phase chilled liquid cooling or desiccant removal as exemplified by the following conditions: 
     EXAMPLE 1 
     A 1,000 scfm stream of air at 95° F. and 100% relative humidity is to be dehumidified to a 34° F. dew point. Such conditions are representative of extreme summer conditions in southern climates. 
     DX Cooling System (“DX”): A DX cooling coil cannot be used as the sole system for such conditions inasmuch as the coil will accumulate ice and airflow will decrease until stopping completely. 
     Chilled Liquid Cooling System (“CL”): It is assumed that the liquid is chilled by a direct expansion system in the primary loop having a saturated suction temp of 20° F. at the heat exchanger. The chilled cooling liquid leaves the heat exchanger at 30° F. The 95° F. entering air leaves the cooling coil at 34° F. Total cooling load would be 228,402 BTU/hr. Refrigeration compressor power consumption according to accepted practice is 27.6 kW. 
     Multiple Stage Dehumidification and Cooling System (“MS”) 
     Stage 1, direct expansion cooling coil saturated suction temp=36° F. at the cooling coil. The 95° F. entering air leaves the expansion cooling coil at 46° F. Subtotal cooling load is 203,306 BTU/hr. Subtotal Refrigeration compressor power consumption is 18.9 kW. 
     Stage 2, the expansion cooling coil in primary loop is operated at a saturated suction temp of 20° F. at the heat exchanger. The chilled cooling liquid leaves the heat exchanger at 30° F. The 46° F. entering air leaves the coil at 34° F. Total cooling load is 25,097 BTU/hr. Subtotal refrigeration compressor power consumption is 3.8 kW. 
     Total cooling=203,306+25,096=228,402 BTU/hr. Total refrigeration compressor power consumption is 3.8+18.9 or 22.7 kW. 
     Comparison: 
     
       
         
               
               
               
               
               
             
               
               
               
               
               
             
           
               
                   
                   
               
               
                   
                 DX 
                 CL 
                 MS 
                 % reduction 
               
               
                   
                   
               
             
             
               
                   
               
             
          
           
               
                 Cooling-Btu/hr 
                 N/A 
                 228,402 
                 228,402 
                 0 
               
               
                 Compressor kW 
                 N/A 
                 27.6 
                 22.7 
                 8.2 
               
               
                 Pumping kW 
                 N/A 
                 1.9 
                 0.2 
                 89.5 
               
               
                 Total KW 
                 N/A 
                 29.8 
                 22.9 
                 30.3 
               
               
                   
               
             
          
         
       
     
     In this example, the chilled water system total kW is 27.9+1.9=29.8 kW. The new Hybrid system hybrid system total kW is 22.7+.2=22.9 kW. The chilled liquid system power consumption is 30.3% higher than the multiple mode system of the invention. 
     EXAMPLE 2 
     A 1,000 scfm stream of air at 95° F. dry bulb and 740 F. at 950 wet bulb to be dehumidified to a 340 F. dew point. This condition is representative of southeastern design conditions. 
     DX cooling system: A DX cooling coil cannot be used inasmuch as the coil will accumulate ice and airflow will decrease until stopping completely. 
     Chilled liquid cooling System: Assume that the liquid is chilled by a direct expansion system in the primary loop having a saturated suction temp of 20° F. at the heat exchanger. The chilled cooling liquid leaves the heat exchanger at 30° F. The 95° 0  F. entering air leaves the cooling coil at 34° F. Total cooling load is 112,089 BTU/hr. Refrigeration compressor power consumption according to accepted practice is 13.7 kW and 1.9 kW pumping power. 
     Desiccant Dehumidifier (“DS”). The desiccant system heats the entering air to 138° F. The regeneration air stream is 148° F. or ten degrees above final supply air temperature. An expansion coil is use to cool the air stream 53° requires 57,240 BTU/hr or 16.8 kW. Post cooling compressor power consumption is 7.0 kW. Total required power for the desiccant system is 23.8 kW. 
     Multiple Stage Dehumidification System 
     Stage 1, direct expansion cooling coil saturated suction temp=36° F. at the cooling coil. The 95° F. entering air leaves the expansion cooling coil at 46° F. Subtotal cooling load is 87,044 BTU/hr. Subtotal Refrigeration compressor power consumption is 7.8 kW kW. 
     Stage 2, Expansion cooling coil in primary loop operates at a saturated suction temp of 20° F. at the heat exchanger. The chilled cooling liquid leaves the heat exchanger at 30° F. The 46° F. entering air leaves the coil at 34° F. Total cooling load=25,045 BTU/hr. Subtotal Refrigeration compressor power consumption=3.2. kW. 
     Total cooling=87,044+25,045=112,089 BTU/hr Total Refrigeration compressor power consumption=11.0 kW. Pumping power is 0.2 kW. 
     Comparison: 
     
       
         
               
               
               
               
             
               
               
               
               
               
             
           
               
                   
                   
               
               
                   
                 Ds 
                 CL 
                 MS 
               
               
                   
                   
               
             
             
               
                   
               
             
          
           
               
                   
                 Cooling BTU/hr 
                 144,172 
                 112,089 
                 112,089 
               
               
                   
                 Power kW 
                 23.8 
                 13.7 
                 11.02 
               
               
                   
                 Pumping KW 
                 0 
                 1.9 
                 0.2 
               
               
                   
                 Total kW 
                 23.8 
                 15.6 
                 11.2 
               
               
                   
                   
               
             
          
         
       
     
     In this example, the chilled water system total kW is 39% higher than the system of the present invention. The desiccant power consumption is 212% higher than the present invention. 
     Other Fluids for the Dehumidification and Cooling System 
     Commonly used refrigeration or heat transfer fluids would be suitable for the secondary liquid system. Some of these include, but are not limited to: glycol solutions, propylene glycol, ethylene glycol, brines, inorganic salt solutions, potassium solutions, potassium formiate, silicone plymers, synthetic organic fluids, eutectic solutions, organic salt solutions, citrus terpenes, hydrofluouroethers, hydrocarbons, chlorine compounds, methanes, ethanes, butane, propanes, pentanes, alcohols, diphenyl oxide, biphenyl oxide, aryl ethers, terphenyls, azeotropic blends, diphenylethane, alkylated aromatics, methyl formate, polydimethylsiloxane, cyclic organic compounds, zerotropic blends, methyl amine, ethyl amine, ammonia, carbon dioxide, hydrogen, helium, water, neon, nitrogen, oxygen, argon, nitrous oxide, sulfur dioxide, vinyl chloride, propylene, R400, R401A, R402B, R401C, R402A, R402B, R403A, R403B, R404A, R405A, R406A, R407A, R407B, R407C, R407D, R408A, R409A, R409B, R410A, R410B, R411A, R411B, R412A, R500, R502, R503, R504, R505, R506, R507A, R508A, R508B, R509A, R600A, R1150, R111, R113, R114, R12, R22, R13, R116, R124, R124A, R125, R134A, R143A, R152A, R170, R610, R611, sulfur compounds, R12B1, R12B 2 , R13B1, R14, R22B1, R23, R32, R41, R114, R1132A, R1141, R1150, R1270, fluorocarbons, carbon dioxide, solutions of water, an d combinations of the above fluids. 
     Other Advantages 
     There is a power cost savings associated with utilizing compressor power consumption to cool a thermal bank on electrical power utility off peak hours. The thermal storage design typically requires a primary refrigeration system operating in a primary loop and carrying a primary refrigerant; a liquid secondary refrigeration system operating in a secondary loop and carrying a secondary liquid refrigerant; heat transfer means for transferring heat from said secondary loop to said primary loop; a secondary cooling coil that is cooled by the secondary loop. In this new invention a similar coil is located in serial air stream association with a first direct expansion cooling coil. With the expenditure of the secondary liquid cooling coil previously financially justified by the serial air stream dehumidification design, thermal energy storage systems enjoy a shorter financial payback time period because the cost of the secondary cooling coil is not applied to the thermal storage system cost. 
     Additionally, this new invention allows the flexibility of operating a cooling or dehumidification system in the primary direct expansion mode, in the secondary liquid cooling mode, or in both modes at once. This feature removes some of the operational risk from thermal energy storage systems by reducing the risk of operational failure during an energy storage capacity failure. This feature removes some of the operational risk from direct expansion primary cooling systems by reducing the risk of operational failure during a compressor failure. 
     Having thus described a presently preferred embodiment of the present invention, it will now be appreciated that the objects of the invention have been fully achieved, and it will be understood by those skilled in the art that many changes in construction and widely differing embodiments and applications of the invention will suggest themselves without departing from the sprit and scope of the present invention. The disclosures and description herein are intended to be illustrative and are not in any sense limiting of the invention, which is defined solely in accordance with the following claims.