Examples of HOSPITAL OPERATING THEATRE Air Handling Unit, Cooling coils

Prof. Dr. Essam E. Khalil Professor of Mechanical Engineering, Mechanical Power Engineering Department Faculty of Engineering, Cairo University, Cairo, Egypt Chairman of the Consulting Engineering Bureau, CEB Chairman of the Egyptian Air Conditioning and Refrigeration Code, HBRC Chairman of the Energy Efficiency of Buildings Code of Egypt EEBC, HBRC

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3.1 Introduction An AHU (Air Handling Unit) is a piece of equipment through which air passes and undergoes different processes in order to deliver air with specific conditions such as; temperature, pressure, moisture content, size of suspended particles and dust, air velocity and air quantity. The AHU has certain components of different arrangement according to the application in hand. But generally, AHU serves one zone with single or multiple spaces. The outer casing may be made of heavy gauge pre-galvanized sheet steel, folded to form sturdy side, top and bottom panels and a self-supporting structure. The supply air ducts may either be directly connected to the unit or may be connected to the unit using a flexible adapter piece, in order to reduce vibration transmission from AHU to air duct.

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3.2 Specifications 3.2.1 Construction Table 0-1 AHU Construction options

Type Description Options 1. Frames Frames are made of high quality

extruded aluminum profile. The frame rigidity is achieved through the use of unique constructions.

2. Panels Panels utilize corrosion resistance galvanized steel that is usually available as double skin panels.

1) Single skin panels 2) Aluminum panels 3) Stainless steel panels 4) epoxy paint for panels 5) Polyester epoxy powder electrostatic paint oven baked panels.

3. Base frame Base frame is constructed from steel channels of heavy gauge galvanized steel depending on the unit size and the number of sections the air handling unit is composed of.

The length and width of the base frame depends on the unit size and length.

4. Access panels and

doors

Double skin door with same panel construction clamped from the construction by four bridge clamps to introduce fully removable access panels furnished with one or two handles for easy release. To achieve air tightness, rubber seal between doors and aluminum frame is provided

1) Access door with inspection window 2) Access door with handles and hinges 3) Access doors with bridge clamps and hinges

5. Insulation For best thermal and acoustical performance, panels and frames are internally insulated with 25 mm thick injected foam (polyurethane) insulation with 40-kg/m- density .

1) 25 mm thick fiberglass with 48 kg/m3 density 2) 50 mm thick fiberglass with 48 kg/m3density 3) rubber insulation

3.2.2 Fan section Fans are used to force the air flow in a determined direction. Fans are either axial (propeller type, tube axial type and vane axial type), or radial (centrifugal). Centrifugal fans (forward, backward and airfoil fans) are belt driven, and must be statically and dynamically balanced.

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Different fan motor arrangements can be provided. Standard fans are forward curved selected for optimum outlet velocities and low sound levels. They may be supplied with a flexible connection between the fan discharge outlet and the unit casing. This will minimize the vibration and accordingly the sound level and completely isolate the fan motor assembly from the rest of the unit structure. Fan selection depends on the application in hand, amount of air needed, and noise level limitations. Options: 1. Inlet guide vanes for backward curved fans and airfoil fans to control air flow rate 2. Belt guard 3. Wire mesh on the fan inlet

Figure 0-1 Fan section Fan motor: The fan motors are mounted on a galvanized steel base which is isolated from the AHU casing with rubber mounts. Motors with 5.5 kW or less are provided with variable pitch pulley for the motor and fixed for the fan. Motors are usually Totally Enclosed Fan Cooled (TEFC). Options: 1) Explosion proof motors 2) Two speed motors 3) Standby motors with manual or automatic change over 4) Spring vibration isolators 5) Frequency inverters 6) Circuit breaker (loose item) 7) External overload (loose item).

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Table 0-2Fan types Fan type Curves & Layout Description 1.Forward curved fan

Runs at a relatively low speed compared to other types for the same capacity. Smaller fan for a given duty, excellent-for fan coil units.

2.Radial fan

Self cleaning. Can be designed for high structural strength to achieve high speed and pressures

3.Backward curved fan

More efficient Power curve has a flat peak so that the motor may be sized to cover the complete range of operation from 0% to 100% air flow for a single speed, non-overloading. Pressure curve is generally steeper than that of the forward curved fan. This results in a smaller change in air volume for any variation in system pressure for selections at comparable percentages of free delivery. Point of maximum efficiency is to the right of the pressure peak, allowing efficient fan selection with a built in pressure reserve. Quieter than other types.

3.2.3 Coils A variety of coils including chilled water, hot water, steam coils and direct expansion coils are available. Coils are designed to deliver their respective duties at optimum performance at all design condition. Coils are manufactured from seamless copper tubes, mechanically expanded into aluminum fins. Coils are usually tested at 3140 kPa air

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pressure in a water bath. They also undergo dry chemical cleaning after coil manufacturing for optimum system cleanliness. Airtight gaskets are used where coil pipes exit the unit casing. The sealing around the coil prevents air by pass. Coils are arranged in staggered or in line form in the direction of airflow. Coils are provided with copper headers within the coil section. They are provided with a manual air vent accessible from outside the casing for quick venting. Coils are available from 1-12 rows for both chilled water and DX- systems and from 1-4 rows for heating coils. Because of the variety of coil input conditions these coils are selected through a computer selection program to match the required conditions. Coil circuiting Water coils can be provided with various types of coil circuiting (half, full and double circuiting) depending on the water flow rate and water pressure drop inside the coil. Direct expansion coils are equipped with suitable size distributor to ensure equal refrigerant fed to all circuits depending on the heat transfer and the refrigerant pressure drop. Other circuiting types can be made when required. Coil connection Coil connections can be provided on either right or left hand side facing air return Options 1) Copper tube/Copper fin coil 2) Protective coating on coil 3) Various outside pipe diameter 4) Expansion valve for direct expansion coils

Figure 0-2 Coil arrangement

Drain pan Drain pan is supplied as standard under the cooling coil. Drain pan is made of 1.5 mm thick painted galvanized steel with a connection from either side. The drain pan is insulated on the sides and underside to prevent condensation. Options: 1) Stainless steel drain pan 2) Double wall insulated drain pan 3) Drain pan connection from both sides

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Figure 0-3 Coil Connections and drain pan

3.2.4 Droplet eliminator To avoid water carry over at high velocity, it is recommended to use a droplet eliminator in the unit. Eliminator blades are manufactured from reinforced polypropylene, encased within a galvanized steel frame, and designed to completely eliminated water carry over from cooling coils with minimal air pressure drop. In most cases droplet eliminators are fitted within the cooling coil module, but droplet eliminator could be fitted anywhere inside the air handling unit if required.

Figure 0-4 Droplet eliminator

3.2.5 Filters Filters are used for trapping undesired objects that may be sucked in to the AHU. The nature and size of these objects may vary from relatively large objects such as; plastic bags, bird feathers, or even whole birds, to relatively small objects such as; specific range of dust particles size. The selection of filters depends on indoor air quality required. Different types of filters and their descriptions are illustrated in the table below.

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Table 0-3 Filter Types Type Material Comment

Flat filter & V-filter Washable aluminum and synthetic.

V-type is used when face velocity is to be reduced.

Bag filter Synthetic media Used when higher level of filtration is required. Efficiency reaching 95%.

HEPA filter High efficiency particulate filters are used when a very high degree of filtration is required such as hospitals. Efficiency reaching 99.99%

Figure 0-5 Air filters

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3.2.6 Mixing Box & Exhaust Box The mixing box module combines fresh air with the circulated return air from the conditioned space. A mixing box may be supplied with fresh and return air dampers, which may or may not be motorized. While the exhaust box function is to exhaust some of the circulated air and return the rest to the supply air stream. Certain filters may be included in the mixing box module. Also an economizer section is optional.

Figure 0-6 Mixing box

Table 0-4 Mixing Box Types Type Function Contents Options

Mixing box Combines fresh air with circulated return air.

Pre-filter and/or bag filter. Manual dampers.

Motorized dampers. Economizer section. Insulation. Exhaust box Exhaust some of the

circulated air and return the rest to the supply air stream.

Manual dampers.

3.2.7 Inlet accessories 1. Insect screen. 2. Rail grill. 3. Rain hood.

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3.2.8 Dampers It is a metallic louver that is mounted on different sections of a duct or AHU and positioned manually or electrically to control the volume of air flowing through this section.

Figure 0-7 Dampers

Table 0-5 Damper Material Options Material options Comment Option

Rigid aluminum frame with multi airfoil blades. Opposed galvanized steel blades.

To reduce pressure drop and sound generated when air passes through the blades. Driven through geared linkage with sealed for life lubrication bearings. Designed for minimized air leakage.

They may be linked for motorized operation. May be supplied with a manually adjustable lever that can be located on either side of the damper.

3.2.9 Sand trap louver Heavy gauge galvanized steel with U-shape plates mounted and encased in a galvanized steel frame. These plates prevent the large particles from entering the AHU with the fresh air, and thus help in the prolonging filter life and the cleanliness of the air stream.

Figure 0-8 Sand tap louver

3.2.10 Sound attenuators Sound attenuators can be supplied in the supply and return airside. These attenuators can be of different lengths depending on the required sound level.

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3.2.11 Heat Recovery sections This custom module includes an energy recovery device that depends on ambient conditions, and uses the exhaust air stream to control the entering out door air conditions. Various types of heat recovery systems are shown in table 3-6. Table 0-6 Types of Heat Recovery Sections

Type Components Theory of operation

Uses

Energy wheel.

A rotating wheel coated with a special material suited for energy transfer.

The supply air flows through one half of the rotary wheel and the exhaust air flows in the counter direction through the other half.

Energy wheel heat exchangers are used in double deck units.

Cross flow heat exchanger.

Consists of small, separated and sealed alternating layers of plates.

Relies on thermal conduction for energy recovery. This type is limited to sensible energy recovery.

Cross flow heat exchanger are used in double deck units

Run around coils.

Consists of two finned-tube coils (air to water heat exchanger) piped together.

One coil in the out door stream and the other in the exhaust air stream.

This type is limited to sensible energy heat transfer.

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3.2.12 Diffuser plate This plate is provided when high face velocities exist, especially when final filters are provided in the AHU. It is made of heavy gauge galvanized steel perforated plate.

Figure 0-9 Diffuser plate

3.2.13 Humidifiers They are used to increase or control the moisture content of the air. Their types vary according to each application. Table 0-7 Humidifiers Theory of Operation

Type Theory of operation Auxiliaries Electric pan humidifier. Air is humidified by

evaporating water in a painted galvanized sheet metal tank using electric element heater.

The humidifier tank is provided with a float valve, drainage output, quick-fill opening and a water level switch.

Steam type. By using immersed electrodes, steam cylinder and stainless steel steam distribution pipe complete with electronic controls for water level regulation and automatic flushing.

Wetted media type. Water is sprayed over the pad area. Air is humidified and cooled while passing through the wetted pad media.

Air washer type. Water droplets are sprayed in the air thus increasing its moisture content until saturation levels if required. The excess water is condensed and collected to be drained away.

-Bolted galvanized sheet metal panels internally sealed for water tightness. -PVC plates specially shaped to for droplet eliminators. -Water sump equipped with the following openings:- Drain connection Supply connection Suction connection Overflow Quick-fill -Inspection window and access door. -Stand pipes with spray nozzles supplying fine water mist.

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3.3 Specific design criteria for Surgery and Critical Care No area of the hospital requires more careful control of the aseptic condition of the environment than does the surgical suite. The systems serving the operating rooms, including cystoscopic and fracture rooms, require careful design to reduce to a minimum the concentration of airborne organisms. The greatest amount of the bacteria found in the operating room comes from the surgical team and is a result of their activities during surgery. During an operation, most members of the surgical team are in the vicinity of the operating table, creating the undesirable situation of concentrating contamination in this highly sensitive area. 3.3.1 Operating Room Studies of operating room air distribution systems and observation of installations in industrial clean rooms indicate that delivery of the air from the ceiling, with a downward movement to several exhaust inlets located on opposite walls, is probably the most effective air movement pattern for maintaining the concentration of contamination at an acceptable level. Completely perforated ceilings, partially perforated ceilings, and ceiling-mounted diffusers have been applied successfully In the average, hospital operating rooms are in use no more than 8 to 12 h per day (excepting emergencies). For this reason and for energy conservation, the air-conditioning system should allow a reduction in the air supplied to some or all of the operating rooms. However, positive space pressure must be maintained at reduced air volumes to ensure sterile conditions. Consultation with the hospital surgical staff will determine the feasibility of providing this feature. A separate air exhaust system or special vacuum system should be provided for the removal of anesthetic trace gases. Medical vacuum systems have been used for removal of nonflammable anesthetic gases. One or more outlets may be located in each operating room to permit connection of the anesthetic machine scavenger hose. Although good results have been reported from air disinfection of operating rooms by irradiation, this method is seldom used. The reluctance to use irradiation may be attributed to the need for special designs for installation, protective measures for patients and personnel, constant monitoring of lamp efficiency, and maintenance. The following conditions are recommended for operating, catheterization, cystoscopic, and fracture rooms: 1. There should be a variable range temperature capability of 20 to 24.5 C. 2. Relative humidity should be kept between 50 and 60%. 3. Air pressure should be maintained positive with respect to any adjoining rooms by supplying 15% excess air. 4. Differential pressure indicating device should be installed to permit air pressure readings in the rooms. Thorough sealing of all wall, ceiling, and floor penetrations and tight-fitting doors are essential to maintaining readable pressure. 5. Humidity indicator and thermometers should be located for easy observation. 6. Filter efficiencies should be in accordance with Table 3-8. 7. Entire installation should conform to the requirements of NFPA Standard 99, Health Care Facilities. 8. All air should be supplied at the ceiling and exhausted or returned from at least two locations near the floor (see Table 3-9 for minimum ventilating rates). Bottom of exhaust outlets should be at least 3 inches above the floor. Supply diffusers should be of the unidirectional type. High-induction ceiling or side wall diffusers should be avoided.

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9. Acoustical materials should not be used as duct linings unless 90% efficient minimum terminal filters are installed downstream of the linings. Internal insulation of terminal units may be encapsulated with approved materials. Duct-mounted sound traps should be of the packless type or have polyester film linings over acoustical fill. 10. Any spray-applied insulation and fireproofing should be treated with fungi growth inhibitor. 11. Sufficient lengths of watertight, drained stainless steel duct should be installed downstream of humidification equipment to assure complete evaporation of water vapor before air is discharged into the room. Control centers that monitor and permit adjustment of temperature, humidity, and air pressure may be located at the surgical supervisors desk. Table 0-8 Filter Efficiencies for central Ventilation and Air conditioning systems in general hospitals Minimum number of filter beds

Area designation Filter efficiencies, %

Filter bed No. 1a

Filter bed No. 2a

Filter bed No. 3b

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Orthopedic operating room Bone marrow transplant operating room Organ transplant operating room

25 90 99.97c

2

General procedure operating rooms Delivery rooms Nurseries Intensive care units Patient care rooms Treatment rooms Diagnostic and related areas

25 90

1 Laboratories Sterile storage

1

Food preparation areas Laundries Administrative areas Bulk storage Soiled holding areas

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Table 0-9 General pressure relationships and ventilation of Surgery and critical care areas Space function Pressure

relationship to adjacent areas

Minimum air change of outdoor air per hour

Minimum total air change per hour

All air exhausted directly to outdoors

Air recirculated within room units

Operating room all outdoor air system

P 15C 15

Yes No

Operating room recirculating air system

P 5 25 Optional No

Delivery room all outdoor air system

P 15 15 Optional No

Delivery room recirculating air system

P 5 25 Optional No

Recovery room E 2 6 Optional No Nursery suite P 5 12 Optional No Trauma room P 5 12 Optional No Anesthesia storage

Optional 8 Yes No

P = Positive N = Negative = Continuous directional control not required e a Ventilation in accordance with ASHRAE Standard 62-1989, Ventilation for Acceptable Indoor Air Quality, should be used for areas for which specific ventilation rates are not given. b Total air changes indicated should be either supplied or, where required, exhausted. c For operating rooms, 100% outside air should be used only when codes require it and only if heat recovery devices are used. d The term trauma room as used here is the first aid room and/or emergency room used for general initial treatment of accident victims. The operating room within the trauma center that is routinely used for emergency surgery should be treated as an operating room. e Although continuous directional control is not required, variations should be minimized, and in no case should a lack of directional control allow the spread of infection from one area to another. Boundaries between functional areas (wards or departments) should have directional control. 3.3.2 Operation theatre air flow The risk of post operative infection is present in all surgical procedures, but can be particularly serious in certain operations, for example, joint replacement. The National Institute of Health (NIH), Office of Research Services, Division of Engineering services, has conducted an extensive study on the issue of operating room ventilation systems and their effect on the protection of the surgical site.

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Several factors can affect postoperative infection, including patient factors, surgical field factors, room factors, and HVAC factors (which is our highlighted concern in this part of the project). The literature agrees that the primary source of bacteria that causes infection are skin scales or particles. These particles are about 10 microns in diameter, and are shed from exposed regions of skin, both from the surgical staff and by the patient. Suggested standards exist for air-conditioning systems for operating theatres in different countries. These standards contain some specific details for the design of the operating room, such as the supply air flow rate. The actual air to be supplied to the room, however, is defined using two factors, which require experimental measurement to be determined. The 1999 ASHRAE Handbook Applications suggests that " the delivery of air from the ceiling, with a downward movement to several exhaust inlets located on opposite walls, is probably the most effective air movement pattern for maintaining the concentration at an acceptable level." The handbook suggests that the temperature range should be between 16.67C and 26.67C, and that positive pressurization should be maintained. It also suggests that the air should be supplied at the ceiling and exhausted or returned from at least two locations near the floor. It suggests that supply diffusers should be of unidirectional type, and that high-induction ceiling or side wall diffusers should be avoided.

Figure 0-10 Typical Operating Theater Set-Up

Generally, the practice of increasing ACH to high levels results in excellent removal of particles via ventilation, but does not necessarily mean that the percentage of particles that strike the surfaces of concern continue to decrease. In a system that provides a laminar flow regime, a mixture of exhaust location levels works better than either low or high level locations only. However, the difference is not significant enough that the low- or high-level location systems are not viable options. Systems that provide laminar flow regimes represent the best option for an operating room in terms of contamination control, as they result in the smallest percentage of particles impacting the surgical site. However, care needs to be taken in the sizing of the laminar flow array. A face velocity of around 30 to 35 fpm (0.15 to 0.18 m/s) is sufficient from the laminar diffuser array, provided that the array size itself is set correctly. To expand on the issue of diffuser array size, it appears that the main factor in the design of the ventilation system is the control of the central region of the operating room. In particular, the operating lights and surgical staff represent a large heat density in the

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middle of the room. Particulates could become caught in buoyant plumes created by these heat-dissipating objects, at which point control of them is lost. However, if a laminar flow type system is employed, the particles are instead driven by the flow to be exhausted. Ideally then, the array size should be large enough to cover the main heat dissipating objects. This is illustrated in Figure 3-11 below.

Figure 0-11 Temperature Distribution inside Operating Theater

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3.3.3 Room pressure for critical environments The method to achieve directional airflow is via the control of the supply and exhaust airflows within and adjacent to the concerned room. 3.3.3.1 Room Pressurization Fundamentals Room pressurization depends on the ability of air to build up within a room. The leakage into or out of room is a key factor. Chapter 26 of the 2001 ASHRAE Handbook- Fundamentals presents a leakage function relationship that correlates a room or building envelope air leakage to the differential pressure producing the flow. ASHRAE defines the leakage function with the presentation of the power law equation as: Q = C (P)n Where: Q is the volumetric rate of flow through an orifice. C is a flow coefficient that depends on the geometry of the orifice. It is empirically determined using a fan pressurization test, similar to the duct leakage test performed by air balancers. DP is the pressure differential across the orifice n is the pressure exponent, commonly around 0.65 per ASHRAE the figure 3-12 below shows the characteristic infiltration curve that represents the power law equation. Thus, if the gaps around a closed door and gaps to adjacent spaces are modeled as an orifice and you know: a) The differential pressure you want to obtain. b) The geometric coefficient of the gaps. c) The empirical exponent n, you can calculate the differential airflow.

Figure 0-12 Infiltration Curve (Power Law Equation)

However, what is the required differential pressure and related differential airflow to contain or keep out contaminants, is our main concern. 3.3.3.2 Recommended Differential Pressure The Centers for Disease Controls Guidelines for Preventing the Transmission of Mycobacterium Tuberculosis in HealthCare Facilities states a minimum differential pressure P of 0.249 Pa is required to achieve a directional airflow into or out of a room. However, this value is challenged as insufficient based on potential thermal stratification in a room, room supply air diffusion, door swings and eddies.

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3.3.3.3 Recommended Differential Airflow The American Conference of Governmental Industrial Hygienists (ACGIH) industrial Ventilation, A Manual of Recommended Practice addresses a quantitative design differential airflow. It states the proper flow differential will depend on the physical condition of the area, but a general guideline would be to set a 5% flow difference but no less than 50 cfm (24 L/s) Standard for Laboratory Ventilation takes a position that controls using room differential air flow set points are preferred over controls that use room differential pressure. A suggested 10% offset between the supply and exhaust airflows and notes this value has no general validity. The text focuses on the containment or exclusion requirements of an open door versus a closed door and the effect on the overall differential airflow to obtain a 50 fpm (0.254 m/s) velocity through an open door. An open door design criteria is impractical considering the volume of the makeup air through the door. Often, most communicating corridors are egress corridors and for smoke control purposes, most building codes prohibit the communicating corridor from providing any significant transfer air to adjacent rooms. Therefore, the high air volume required to contain or keep out contaminants through an open door would violate the code. The use of an air lock is suggested for critical applications, thereby obviating the potential for a continuous open door path from the room to the communication corridor. 3.3.3.4 Room Pressure for a Protective Environment The airlock protects a contained area against building pressure fluctuations. In a hospital patients rooms are recommended to have the following; The walls and floor penetrations are to be sealed to best of general construction standards that can be from excellent to sufficient. Two access panels, later sealed, penetrate the corridor-to-room wall above the ceiling. The toilet exhaust should be common to other toilet exhausts. A special sink in the patient room (not bathroom) had an open gap drain to another space below for sanitation purposes. The room supply should be via a pressure independent primary air HEPA filtered fan-powered series box. The return to the fan-powered box may be in the room. The house exhaust for the room is served by a pressure independent exhaust box. The supply and exhaust airflows are to be measured with an air volume hood. The differential pressures are measured by placing the static probe in the middle of the room, routing the tube through the door undercut and connecting the probe to the digital manometer out side the room. An airflow direction indicator should be placed above the entry door to show whether the room was under positive pressure.

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3.3.3.5 Door Swings and Anterooms For a positive pressure room it is recommended that the entry doors are to be sealed by a gasket, with sliding break-away doors. If a standard swing door is used, it is recommended to swing out of the room for a negative room and swing into the room for a positive room. This may not always be practical. For example, hospital isolation room doors that are located off the main corridor cannot swing out into the corridor. In such cases, it is advised to use an air lock. An airlock (anteroom) should be used whenever possible. The anteroom traps any escaped air from a negative room and isolates corridor air from a positive room. Because the anteroom is a trap, it should incorporate a high air change rate of around 12 ACH or higher and the differential cfm should be zero or neutral to allow overall desired directional airflow between the corridor and the concerned room. 3.3.3.6 Comments Some basic points for designing for proper room pressurization based on differential airflow settings include: Seal the room. Meet or exceed minimum codes for air change rates. Incorporate industry regulations and practice for minimum air change rates and room pressure. However, as a minimum, strive for 2.49 Pa to 12.45 Pa differential pressure. When designing the HVAC system to obtain the desired room pressurization/directional air flow for 200 ft2 (18.58 m2) rooms, consider the following points: Rooms should have a minimum negative or positive pressure of 2.49 Pa where 12.45 Pa or higher is preferred. Codes and industry regulations and practice may dictate specific limits. Air balancer specs for positive rooms should be considered. For negative rooms, the makeup air should be provided via a supply outside the room. For positive rooms, exfiltration of air should be accommodated by an exhaust outside the room. All room penetrations above and below the ceiling and the ductwork should be well sealed. The ceiling should be tight as possible, preferably sheetrock or concrete deck. Specify surface mount or recessed vapor-tight, or non-re turn-air light fixtures. Each entry door to the room should be sealed on its top and sides (including astragal vertical joint seal for leaf or double doors) and include an adjustable bottom seal. A sliding entry door is preferred over a swing door. If a swing door is used, it should open out of a negative room or open into a positive room. Anterooms should be used whenever possible with 12 air changes per hour (ACH) minimum (codes and industry regulations and practice may dictate higher values) and a neutral pressure where the supply and exhaust airflow quantities are equal. An airflow direction indicator should be installed to visually see the dynamics of the room pressurization.

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3.4 Example of Operating Theater Load Estimation 3.4.1 General Zone Data Typical operating rooms arrangement is shown in appendix D, the largest room was found to be 52 m2 area with a ceiling height of 3 m. There are no walls facing outside areas. The partitions to adjacent operating rooms should represent no load on the air conditioning facility. While the partitions to corridors and antiseptic storage represent a load due to temperature variations with respect to the operating room. Also the ceiling and floor will participate in the load due to difference in temperatures. Floor area : 52 m2 Building weight : Medium Lighting Fluorescent light : 640 watts Fixture type : Rec., Not vented Operating theatre lighting fixture : 100 Watts Total watts : 740 watts Wattage Multiplier. : 1.25 Total wattage : 925 watts Other Electric Electrocardiogram (monitor) : 130 watts X-ray screen: 200 Watts Pulse measuring machine: 160 Watts Operation spot light: 2500 Watts Dialysis machine : 480 Watts Portable sterilization machine : 670 Watts Ultrasonography: 110 Watts Computer and monitors for laser and endoscope systems: 325 Watts DC shock: 200 Watts Total wattage: 4450 Watts People Number of people : 10 people Activity level : medium work Sensible gain : 86.5 W/Person Latent gain : 133.3 W/Person Miscellaneous loads Sensible : 1000 Watts Latent: : 500 Watts Reheater Power: 8131 Watts Required room conditions Dry bulb temperature: 21C Relative humidity: 50 % Weather data The weather data collected by the meteorological authority in Egypt as shown in appendix A state that the extreme weather conditions occurs in July with 43.33 C dry bulb temperature and a 27.66 % relative humidity with air enthalpy of 83.35 kJ/kg this condition is extreme in the sense of dry bulb temperature while if revising the July

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average weather condition it is shown that the air enthalpy is as high as 86.89 kJ/kg. This leads us to take the July average conditions to run the load calculation on the operating room. But as stated in the AIAA paper (AIAA 4199 - 2003), By Dr. E. E. Khalil, the weather extreme conditions should be modified to 40 C dry bulb temperature and 50 % relative humidity, in order to account for global warming as the weather conditions in Cairo deteriorates each year. Thus we have concluded the recommended weather conditions in Appendix B. As such the weather conditions taken in the cooling coil design are; Dry bulb temperature: 40C Relative humidity: 50 % Air flow rate As stated before, the dominating factor in the design of Operating rooms air conditioning is the air quality. The air quality is dominated by the air change per hour. The minimum air change per hour advised by the ASHRAE is 15 ACH per hour for operating rooms with all outdoor air. For extra safety we have chosen it to be 25 air changes per hour. The 25 ACH per hour will lead to an air flow rate of 1.085 m3/s. The air velocity should be kept at the lower boundary in order to avoid water drift with air flow. The air velocity should be within the range from 1 to 3 m/s. Refrigerant Choosing chilled water system for cooling, the chilled water would be the refrigerant. Refrigerant: Chilled water Inlet conditions: 6C outlet conditions: 12C The water velocity inside tubes should be in the range of 0.5 to 1.5 m/s 3.4.2 Room sensible load: Solar heat gain There is no solar heat gain through glass or fenestration as there are no exterior walls. As well there is no solar heat gain through walls . Also assuming there is a conditioned space above, there would be no solar heat gain through the roof.

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Transmission gain except walls and roof The adjacent room air conditions are so close to the inside room conditions and the variation are so small and could not be computed as the CLTD (cooling load temperature difference) factor is zero referring to the CARRIER HANDBOOK for load estimation. Internal heat People: 7 * 86.5 = 605.5W Electric equipment: 4450 Watts Lights: 740 * 1.25 = 925 W Miscellaneous load 1000 W Total room sensible heat ( RSH): RSH = 605.5 + 4450 + 925 + 1000 = 6980.5 W 3.4.3 Room latent load: Infiltration The infiltration doesn't exist as the operating theater should be positively pressurized to maintain higher indoor air quality. That is why the infiltration would not be accounted for in the room latent load. People 7 * 133.3 = 933.1 W Steam and other appliances There are no appliances that would generate steam or any source of latent load inside the operating theater. Miscellaneous load 500 W Total room latent heat (RLH): RLH = 933.1 + 500 = 1433.1 W 3.4.4 Room sensible heat factor (RSHF) RSHF = (6980.5)/(6980.5 + 1433.1) = 0.83 3.4.5 Supply air temperature But the supply air temperature is constrained by the following equation:

airpair tCVRSH D=r

thus the supply air temperature would be,

tc = tr 02.1085.12.15.6980

= pair CV

RSHr = 21-5.25 = 15.74 C

3.4.6 Psychrometric Processes

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Figure 0-13 Psychrometric Represaentation of Load Estimation Investigating the psychrometric chart shown in figure 3-13, we have found the following: Air mass flow rate 1.3 kg/s On coil conditions: 40 C dbt and 50 % RH. Off coil conditions: 9.41 C dbt and 99.41 % RH. 3.4.7 Results Room effect Supply air conditions: 15.56 C dbt and 66.37 % relative humidity. Room air conditions: 21 C dbt and 50% relative humidity. Room sensible heat: 6.98 kW. Room latent heat: 1.433 kW. Air conditioning apparatus Total cooling capacity: 92kW. Moisture removal: 20.4 g/s. On coil air conditions: 40C dbt and 50% relative humidity. Off coil air conditions: 9.3C dbt and 99.4% relative humidity.

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3.4.8 Computer assisted load estimation programs: Using Carrier load estimation program, the E20.II as shown in appendix C, the load came to be about 93 kW and the leaving dry bulb, wet bulb was 9.3/9.3 C. While when using advanced versions of load estimation programs the load was 94 kW and the dry bulb/wet bulb temperatures were 9.9/9.9 C, when a return air plenum was installed. When using the same advanced software but without the installation of a return air plenum the load calculated was 109.6 kW and the dry bulb/wet bulb temperatures were 9.9/9.9 C, as shown in appendix D and E. The existence of return air plenum reduces the coil load greatly as the return air is not exhausted directly to the atmosphere, but rather passed through a space above the conditioned room to remove first some of the load. When the air passes through the return air plenum it carries some of the lighting load (in case of recessed not vented fixtures), as well as part of the external load.

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3.5 Example of Cooling Coil Design 3.5.1 Theoretical background The cooling coil is the major part of the air handling unit as it is responsible of cooling air. When we talk about air treatment we happen to stress more on the cooling coil and methods of reducing the air's temperature due to the nature of the region we are living in. The majority of the equipment used today for cooling and dehumidifying an airstream under forced convection incorporates a coil section that contains one or more cooling coils assembled in a coil bank arrangement. Such coil sections are used extensively as components in room terminal units; larger factory-assembled, self-contained air conditioners; central station air handlers; and field built-up systems. The applications of each type of coil are limited to the field within which the coil is rated. Other limitations are imposed by code requirements, proper choice of materials for the fluids used, the configuration of the air handler, and economic analysis of the possible alternatives for each installation. 3.5.1.1 Coil Construction and Arrangement In finned coils, the external surface of the tubes is the primary surface, and the fin surface is the secondary surface. The primary surface generally consists of rows of round tubes or pipes that may be staggered or placed in line with respect to the airflow. Flattened tubes or tubes with other nonround internal passageways are sometimes used. The inside surface of the tubes is usually smooth and plain, but some coil designs have various forms of internal fins or turbulence promoters to enhance performance. The individual tube passes in a coil are usually interconnected by return bends to form the serpentine arrangement of multipass tube circuits. Coils are usually available with different circuit arrangements and combinations offering varying numbers of parallel water flow passes within the tube core as shown in Figure 3-14. Cooling coils of water, aqueous glycol, brine, or halocarbon refrigerants usually have aluminum fins on copper tubes, although copper fins on copper tubes and aluminum fins on aluminum tubes (excluding water) are also used. Adhesives are sometimes used to bond header connections, return bends, and fin-tube joints, particularly for aluminum-to-aluminum joints. Certain special-application coils feature an all-aluminum extruded tube-and-fin surface. Common core tubes outside diameters are 5/16, 3/8, 1/2, 5/8, 3/4, and 1 inch, with fins spaced 4 to 18 per inch. Tube spacing ranges from 0.6 to 3.0 inch on equilateral (staggered) or rectangular (in line) centers, depending on the width of individual fins and on other performance considerations. Fins should be spaced according to the job to be performed, with special attention given to air friction; possibility of lint accumulation; and frost accumulation, especially at lower temperatures. Tube wall thickness and the required use of alloys other than copper are determined mainly by the coils working pressure and safety factor for hydrostatic pressure. Fin-type and header construction also play a large part in this determination. 3.5.1.2 Water Coils Good performance of water-type coils requires both the elimination of all air and water traps within the water circuit and the proper distribution of water. Unless properly vented, air may accumulate in the coil tube circuits, reducing thermal performance and possibly causing noise or vibration in the piping system. Air vent and drain connections are

26

usually provided on the coil water headers, but this does not eliminate the need to install, operate, and maintain the coil tube core in a level position. Individual coil vents and drain plugs are often incorporated on the headers (Figure 3-14).

Figure 0-14 Typical water circuit arrangement

Depending on performance requirements, the water velocity inside the tubes usually ranges from approximately 0.3 to 2.5 m/s, and the design water pressure drop across the coils varies from about 1.5 to 15 m of water head. The core tubes of properly designed and installed coils should feature circuits that 1. Have equally developed line length; 2. Are self-draining by means of gravity during the coils off cycle; 3. Have the minimum pressure drop to aid in water distribution from the supply header without requiring an excessive pumping head; 4. Have equal feed and return by the supply and return header. Design for the proper in-tube water velocity determines the circuitry style required. Multirow coils are usually circuited to the cross-counterflow arrangement and oriented for top-outlet/ bottom-feed connection. 3.5.1.3 Flow Arrangement In the air-conditioning process, the relation of the fluid flow arrangement within the coil tubes to the coil depth greatly influences the performance of the heat transfer surface. Generally, air-cooling and dehumidifying coils are multirow and circuited for counterflow arrangement. The inlet air is applied at right angles to the coil s tube face (coil height), which is also at the coils outlet header location. The air exits at the opposite face (side) of the coil where the corresponding inlet header is located. Counterflow can produce the highest possible heat exchange within the shortest possible (coil row) depth because it has the closest temperature relationships between tube fluid and air at each side of the coil. 3.5.1.4 Applications Figure 0-15 shows a typical arrangement of coils in a field built-up central station system. All air should be filtered to prevent dirt, insects, and foreign matter from accumulating on the coils. The cooling coil (and humidifier, when used) should include a drain pan under

27

each coil to catch the condensate formed during the cooling cycle (and the excess water from the humidifier). The drain connection should be on the downstream side of the coils, be of sufficient size, have accessible cleanouts, and discharge to an indirect waste or storm sewer. The drain also requires a deep-seal trap so that no sewer gas can enter the system. Precautions must be taken if there is a possibility that the drain might freeze. The drain pan, unit casing, and water piping should be insulated to prevent sweating. Factory-assembled central station air handlers incorporate most of the design features outlined for field built-up systems. These packaged units can generally accommodate various sizes, types, and row depths of cooling and heating coils to meet most job requirements. This usually eliminates the need for field built-up central systems, except on very large jobs.

Figure 0-15 Cooling Coil assembly inside Typical Application

The design features of the coil (fin spacing, tube spacing, face height, type of fins), together with the amount of moisture on the coil and the degree of surface cleanliness, determines the air velocity. Generally, condensate water begins to be blown off a plate fin coil face at air velocities above 3 m/s. Water blow-off from the coils into air ductwork external to the air-conditioning unit should be prevented. However, water blow-off from the coils is not usually a problem if coil fin heights are limited to 1.1 m. and the unit is set up to catch and dispose of the condensate. When a number of coils are stacked one above another, the condensate is carried into the airstream as it drips from one coil to the next. A downstream eliminator section could prevent this, but an intermediate drain pan and/or condensate trough to collect the condensate and conduct it directly to the main drain pan is preferred. Extending downstream of the coil, each drain pan length should be at least one-half the coil height, and somewhat greater when coil airflow face velocities and/or humidity levels are higher. When water is likely to carry over from the air-conditioning unit into external air ductwork, and no other means of prevention is provided, eliminator plates should be installed on the downstream side of the coils.

28

Air-cooling and dehumidifying coil frames, as well as all drain pans and troughs, should be of an acceptable corrosion-resistant material suitable for the system and its expected useful service life. The air handlers coil section enclosure should be corrosion-resistant; be properly double-wall insulated; and have adequate access doors for changing air filters, cleaning coils, adjusting flow control valves, and maintaining motors. 3.5.1.5 COIL SELECTION When selecting a coil, the following factors should be considered: Job requirements; cooling, dehumidifying, and the capacity required to properly balance with other system components Temperature conditions of entering air Available cooling media and operating temperatures Space and dimensional limitations Air and cooling fluid quantities, including distribution and limitations Allowable frictional resistances in air circuit (including coils) Allowable frictional resistances in cooling media piping system (including coils) Characteristics of individual coil designs and circuitry possibilities Individual installation requirements such as type of automatic control to be used; presence of corrosive atmosphere; design pressures; and durability of tube, fins, and frame material. Air quantity is affected by such factors as design parameters, codes, space, and equipment. The resistance through the air circuit influences the fan power and speed. This resistance may be limited to allow the use of a given size fan motor, to keep the operating expense low, or because of sound level requirements. The air friction loss across the cooling coil; in summation with other series air pressure drops for such elements as air filters, water sprays, heating coils, air grilles, and ductwork; determines the static pressure requirement for the complete airway system. The static pressure requirement is used in selecting the fans and drives to obtain the design air quantity under operating conditions. The conditioned air face velocity is determined by economic evaluation of initial and operating costs for the complete installation as influenced by 1. Heat transfer performance of the specific coil surface type for various combinations of face areas and row depths as a function of the air velocity; 2. Air-side frictional resistance for the complete air circuit (including coils), which affects fan size, power, and sound-level requirements; 3. Condensate water carryover considerations. The allowable friction through the water or brine coil circuitry may be dictated by the head available from a given size pump and pump motor, as well as the same economic factors governing the air side made applicable to the water side. Additionally, the adverse effect of high cooling water velocities on erosion-corrosion of tube walls is a major factor in sizing and circuitry to keep tube velocity below the recommended maximums. On larger coils, water pressure drop limits of 4.5 to 6 m water usually keep such velocities within acceptable limits of 0.183 to 0.366 m/s, depending on circuitry design. Coil ratings are based on a uniform velocity. Design interference with uniform airflow through the coil makes predicting coil performance difficult as well as inaccurate. Such airflow interference may be caused by the entrance of air at odd angles or by the inadvertent blocking of a portion of the coil face. To obtain rated performance, the

29

volumetric airflow quantity must be adjusted on the job to correspond to that at which the coil was rated and must be kept at that value. In the case of dehumidifying coils, it is important that the proper amount of surface area be installed to obtain the ratio of air-side sensible-to-total heat required for maintaining the air dry-bulb and wet-bulb temperatures in the conditioned space. This is an important consideration when preconditioning is done by reheat arrangement. The same room air conditions can be maintained with different air quantities (including outside and return air) through a coil. However, for a given total air quantity with fixed percentages of outside and return air, there is only one set of air conditions leaving the coil that will precisely maintain the room design air conditions. Once the air quantity and leaving air conditions at the coil have been selected, there is usually only one combination of face area, row depth, and air face velocity for a given coil surface that will precisely maintain the required room ambient conditions. Therefore, in making final coil selections it is necessary to recheck the initial selection to ensure that the leaving air conditions, as calculated by a coil selection computer program or other procedure, will match those determined from the cooling load estimate. Coil ratings and selections can be obtained from manufacturers catalogs. Most catalogs contain extensive tables giving the performance of coils at various air and water velocities and entering humidity and temperatures. Most manufacturers provide computerized coil selection programs to potential customers. The final choice can then be made based on system performance and economic requirements. 3.5.1.6 AIRFLOW RESISTANCE A cooling coils airflow resistance (air friction) depends on the tube pattern and fin geometry (tube size and spacing, fin configuration, and number of in-line or staggered rows), the coil face velocity, and the amount of moisture on the coil. The coil air friction may also be affected by the degree of aerodynamic cleanliness of the coil core; burrs on fin edges may increase coil friction and increase the tendency to pocket dirt or lint on the faces. A completely dry coil, removing only sensible heat, offers approximately one-third less resistance to airflow than a dehumidifying coil removing both sensible and latent heat. For a given surface and airflow, an increase in the number of rows or number of fins increases the airflow resistance. Therefore, the final selection involves the economic balancing of the initial cost of the coil against the operating costs of the coil geometry combinations available to adequately meet the performance requirements. The aluminum fin surfaces of new dehumidifying coils tend to inhibit condensate sheeting action until they have aged for a year. Recently developed hydrophilic aluminum fin surface coatings reduce the water droplet surface tension, producing a more evenly dispersed wetted surface action at initial start-up. Manufacturers have tried different methods of applying such coatings, including dipping the coil into a tank, coating the fin stock material, or subjecting the material to a chemical etching process. Tests have shown as much as a 30% reduction in air pressure drop across a hydrophilic coil as opposed a new untreated coil. 3.5.1.7 HEAT TRANSFER The heat transmission rate of air passing over a clean tube (with or without extended surface) to a fluid flowing within it is impeded principally by three thermal resistances.

30

The first, from the air to the surface of the exterior fin and tube assembly, is known as the surface air-side film thermal resistance. The second is the metal thermal resistance to the conductance of heat through the exterior fin and tube assembly. The third is the in-tube fluid-side film thermal resistance, which impedes the flow of heat between the internal surface of the metal and the fluid flowing within the tube. For some applications, an additional thermal resistance is factored in to account for external and/or internal surface fouling. Usually, the combination of the metal and tube-side film resistance is considerably lower than the air-side surface resistance. For a reduction in thermal resistance, the fin surface is fabricated with die-formed corrugations instead of the traditional flat design. At low airflows or wide fin spacing, the air-side transfer coefficient is virtually the same for flat and corrugated fins. Under normal comfort conditioning operation, the corrugated fin surface is designed to reduce the boundary air film thickness by undulation of the passing airstream within the coil; this produces a marked improvement in heat transfer without much airflow penalty. Further fin enhancements, including the louvered and lanced fin designs, have been driven by the desire to duplicate throughout the coil depth the thin boundary air film characteristic of the fins leading edge. Louvered fin design maximizes the number of fin surface leading edges throughout the entire secondary surface area and increases the external secondary surface area, as through the multiplicity of edges. The transfer of heat between the cooling medium and the airstream across a coil is influenced by the following variables: Temperature difference between fluids Design and surface arrangement of the coil Velocity and character of the airstream Velocity and character of the in-tube coolant With water coils, only the water temperature rises. With coils of volatile refrigerants, an appreciable pressure drop and a corresponding change in evaporating temperature through the refrigerant circuit often occur. The rating of direct-expansion coils is further complicated by the refrigerant evaporating in part of the circuit and superheating in the remainder.

31

3.5.1.8 PERFORMANCE OF SENSIBLE COOLING COILS The performance of sensible cooling coils depends on the following factors. 1. The overall coefficient Uo of sensible heat transfer between airstream and coolant fluid 2. The mean temperature difference tm between airstream and coolant fluid 3. The physical dimensions of and data for the coil (such as coil face area Aface and total external surface area Ao ) with characteristics of the heat transfer surface The sensible heat cooling capacity q s of a given coil is expressed by the following equation: qs = Uo FsAface Nrtm (1a) with Fs = Ao /(Aface * Nr) (1b) Assuming no irrelevant heat losses, the same amount of sensible heat is lost from the airstream:

)( 1aaopaira irs ttCmq -= (2a)

with

oairfacefaceairAVm r**= (2b)

The same amount of sensible heat is absorbed by the coolant; for a nonvolatile type, it is )(* inroutrwaterwater ttCmqs -*=

(3) For a nonvolatile coolant in thermal counterflow with the air, the mean temperature difference in Equation (1a) is expressed as

--

---=D

inra

outrao

inraoutraom

tttt

ttttt

1

1

(ln

)()( (4)

These calculations are based on various assumptions; among them that U for the total external surface is constant. While this assumption is generally not valid for multirow coils, the use of cross-flow temperature differences is preferable to Equation (4), which applies only to counterflow. However, the use of the log mean temperature difference is widespread. The overall heat transfer coefficient Uo for a given coil design, whether bare-pipe or finned-type, with clean, non-fouled surfaces, consists of the combined effect of three individual heat transfer coefficients: 1. The film coefficient hc of sensible heat transfer between air and the external surface of the coil 2. The unit conductance 1/R md of the coil material (i.e., tube wall, fins, etc.) 3. The film coefficient hr of heat transfer between the internal coil surface and the coolant fluid within the coil. For a bare-pipe coil, the overall coefficient of heat transfer for sensible cooling (without dehumidification) can be expressed by a simplified basic equation:

)/(2/)()/1(1

roiioco hAkDDh

U+-+

= (5a)

When pipe or tube walls are thin and made of material with high conductivity (as in typical heating and cooling coils), the term (Do - Di)/2k in Equation (5a) frequently becomes negligible and is generally disregarded. (This effect in typical bare-pipe cooling

32

coils seldom exceeds 1 to 2% of the overall coefficient.) Thus, the overall coefficient for bare pipe in its simplest form is

)/()/1(1

roic hAhUo

+= (5b)

For finned coils, the equation for the overall coefficient of heat transfer can be written

)/()/1(1

roic hAhUo

+=

h (5c)

Where the fin effectiveness allows for the resistance to heat flow encountered in the fins. It is defined as

ops AAEA /)( +=h (6) Where E is the fin efficiency. For typical cooling surface designs, the surface ratio Aoi ranges from about 1.03 to 1.15 for bare pipe coils and from 10 to 30 for finned coils. Estimation of the air-side heat transfer coefficient hc is more difficult because well-verified general predictive techniques are not available. Hence, direct use of experimental data is usually necessary. With a given design and arrangement of heat transfer surface used as cooling coil core material for which basic physical and heat transfer data are available to determine Uo from Equation (5a), Equation (5b), and Equation (5c), the selection, sizing, and performance calculation of sensible cooling coils for a particular application generally reduces to the heat transfer surface area Ao or the coil row depth Nr for a specific coil size is required and initially unknown. The sensible cooling capacity q s, flow rates for both the air and the coolant, entrance and exit temperatures of both fluids, and mean temperature difference between fluids are initially known or can be assumed or determined from Equation (2a), Equation (3), and Equation (4). The required surface area Ao or coil row depth Nr can then be calculated directly from Equation (1a). 3.5.1.9 PERFORMANCE OF DEHUMIDIFYING COILS A dehumidifying coil normally removes both moisture and sensible heat from entering air. In most air conditioning processes, the air to be cooled is a mixture of water vapor and dry air gases. Both lose sensible heat when in contact with a surface cooler than the air. The removal of latent heat through condensation occurs only on the portions of the coil where the surface temperature is lower than the dew point of the air passing over it. As the leaving dry-bulb temperature is lowered below the entering dew-point temperature, the difference between the leaving dry-bulb temperature and the leaving dew point for a given coil, airflow, and entering air condition is lowered. When the coil starts to remove moisture, the cooling surfaces carry both the sensible and latent heat load. As the air approaches saturation, each degree of sensible cooling is nearly matched by a corresponding degree of dew-point decrease. The latent heat removal per degree of dew-point change is considerably greater. The potential or driving force for transferring total heat qt from the airstream to the tube-side coolant is composed of two components in series heat flow: (1) An air-to-surface air enthalpy difference (ha - hi) (2) A surface-to-coolant temperature difference (ti - tr). Figure 3-16 is a typical thermal diagram for a coil in which the air and a nonvolatile coolant are arranged in counterflow. The top and bottom lines in the diagram indicate,

33

respectively, changes across the coil in the airstream enthalpy ha and the coolant temperature tr. To illustrate continuity, the single middle line in Figure 3-16 represents both surface temperature ti and the corresponding saturated air enthalpy h i, although the temperature and air enthalpy scales do not actually coincide as shown. The differential surface area dAw represents any specific location within the coil thermal diagram where operating conditions are such that the air-surface interface temperature ti is lower than the local air dew-point temperature. Under these conditions, both sensible heat and latent heat are removed from the airstream, and the cooler surface actively condenses water vapor.

Figure 0-16 Two-component Driving Force Between Dehumidifying Air and Coolant

Neglecting the enthalpy of the condensed water vapor leaving the surface and any radiation and convection losses, the total heat lost from the airstream in flowing over dAw is

aairt dhmdq *-= (7)

This same total heat is transferred from the airstream to the surface interface through both sensible and latent processes.

)(** iaocs ttdAhdq -=

fgcondl hdmdq *.= )(**. iaoDcond dAhdm yy -=

fgiaoDiaoct hdAhttdAhdq *))(**()(** yy -+-=\ (8) but for a Lewis number of unity the following is valid

)_(/ airwetpcD Chh = and moist air enthalpy could be calculated as follows

))(*)_((** titavaporwaterCphfgataCpha -++= y

)_(

**)(

airwetp

coiat C

hdAhhdq

-=\ (9)

hi ti

hi

ti hi

34

The total heat transferred from the air-surface interface across the surface elements and into the coolant is equal to that given in Equation (7) and Equation (9):

)(** riirt ttdAhdq -= (10) The same quantity of total heat is also gained by the chilled water in passing across dAw

)(* rwaterwatert dtCmdq-= (11)

If Equation (9) and Equation (10) are equated and the terms rearranged, an expression for the coil characteristic Rcf is obtained:

ia

ri

irp

occf hh

ttdAhC

dAhR--

== (12)

Equation (12) shows the basic relationship of the two components of the driving force between air and coolant in terms of two principal heat transfer coefficients. For a given coil, these tow heat transfer coefficient of air, and combined metal in-tube fluid can be determined for the particular application, which gives a fixed value for Rcf. Equation (12) can then be used to determine point conditions for the interrelated values of airstream s enthalpy ha coolant temperature tr; surface temperature ti and the enthalpy hi of saturated air corresponding to the surface temperature. When both t i and hi are unknown, a trial-and-error solution is necessary. Figure 3-17 shows a typical thermal diagram for a portion of the coil surface when it is operating dry. The illustration is for counter flow with chilled water as a refrigerant. The diagram at the top of the figure 3-17 illustrates a typical coil installation in an air duct with tube passes circuited countercurrent to airflow. Locations of the entering and leaving boundary conditions for both air and coolant are shown. The thermal diagram in Figure 3-17 is of the same type as that in Figure 3-16, showing three lines to illustrate local conditions for the air, surface, and coolant throughout a coil. The dry-wet boundary conditions are located where the coil surface temperature t ib equals the entering air dew-point temperature dpt o; Thus, the surface area Ad to the left of this boundary is dry, with the remainder Aw of the coil surface area operating wet. When using fluids or halocarbon refrigerants in a thermal counter flow arrangement as illustrated in Figure 3-17, the dry-wet boundary conditions can be determined from the following relationships:

waterwater

air

aao

inroutr

Cmm

hhtt

y

=--

=

1

(13)

yRhRhytdpt

hcf

Dptcfaoroutoab

o

+

++-= (14)

35

Figure 0-17 Thermal Diagram for General Case When Coil Surface Operates Partially Dry The value of hab from Equation (14) serves as an index of whether the coil surface is operating fully wetted, partially dry, or completely dry, according to the following three limits: If hab hao the surface is fully wetted. If hao > hab > ha1, the surface is partially dry. If hab ha1, the surface is completely dry. Other dry-wet boundary properties are then determined:

oib Dptt = (15)

pabaoaoab chhtt /)( --= (16) )( abaoproutrb ttcytt --= (17)

The dry surface area Ad required and capacity qtd are calculated by conventional sensible heat transfer relationships, as stated before in section 6.5.1.8.

)/()/1(1

roic hAhUo

+=

h (18)

With ops AAEA /)( +=h (19)

The mean difference between air dry bulb temperature and coolant temperature, using symbols from Figure 3-18, is

( ) ( )

( ) ( )[ ]rbabroaorbabaao

m tttttttt

t-----

=D/ln

1 (20)

36

The dry surface area required is

mo

tdd tU

qAD

= (21)

The air-side total heat capacity is ( )abaopatd ttcmq -= (22a) From the coolant side, ( )rbroutwaterwatertd ttcmq -= (22b) The wet surface area Aw and capacity qtw are determined by the following relationships, using terminology in Figure 3-18. The air-side total heat capacity is

( )[ ]fwaaoatw hhhmq +-= 1 (23a) The enthalpy hfw of condensate removed is

( ) 221 apwfw tch yy -= (23b) Note that hfw for normal air-conditioning applications is about 0.5% of the air stream enthalpy difference (hao ha1) and is usually neglected. The coolant-side heat capacity is ( )rinrbrwatertw ttcmq -= (23c)

Figure 0-18 Thermal Diagram for General Case When Coil Surface Operates Partially Dry The coil surface is divided into n segments, which results into ( n + 1 ) station. The heat transfer through each element can be described as follows:

++

+

=-=

++

++

+p

jijijaja

jjcjajaairii C

hhhh

Ahhhmq22

*)(

)1(,,)1(,,

1)1(,,1 (24)

For each element, we may assume constant heat transfer rate which is equal to the total heat transfer rate divided by the number of elements. This assumption will allow us to

1 2 3 3 4 n n+1

Condensate

Air flow

Chilled water flow

Surface temperature ti

37

calculate the air enthalpy, refrigerant temperature, surface temperature, and enthalpy of air at surface temperature. Thus the element area can be calculated as:

Cphhhh

h

hhmA

jijijajac

jajaairjj

/22

*

)(

)1(,,)1(,,

)1(,,1

+-

+

-=

++

+

+ (25)

then the total required outside surface area required is

The total surface area requirement of the coil is Ao = Ad+Aw . (26) The total heat capacity for the coil is qt = qtd + qtw (27) Now it is required to check the value of the off coil dry bulb temperature, this is done using the following relations:

--

-=-= ++++

+ 22

**)(** )1(,,)1(,,1)1(,,1,jijijaja

jjcjajapairjjs

ttttAhttCmq (28)

thus the dry bulb temperature is calculated at each station until we reach the final stage. The exit dry bulb temperature should satisfy the required design, if not some of the assumed parameters during design which would affect the convection heat transfer coefficients and the surface temperature. 3.5.2 Cooling coil design strategy: The cooling coil design should be designed based on the following conditions: 1. Air inlet conditions. 2. Air outlet conditions. 3. Air flow rate. The available data for the Operating theater were: 1. Air inlet conditions. Dry bulb temperature: 40 C. Relative Humidity : 50 % 2. Air outlet conditions. Dry bulb temperature: 9.3 C. Relative Humidity : 99.41 % 3. Air flow rate. 1.085 m3/s. Now we have to choose the coil configuration from the manufacturer data which are summarized in Table 3-10 and attached figure 3-19.

Table 0-10 Surface Area Data Data Surface 1 Surface 2 Dimensions, (nomenclature according to figure 3-19) A, tube outside diameter, mm 10.2108 17.1704 B, tube spacing across face, mm 25.4 381 C, tube spacing between rows, mm 22 44.45

1+= jjo AA

38

D, spacing of fins, center to center, mm 3.175 3.2766 E, thickness of aluminum fins,mm 0.3302 0.4064 Flow passage hydraulic diameter, 4rh (Dh) 0.302768 0.322072 Area Data Fs, External surface area /(Face area) (No. of rows) 12.92 22.86 Aoi,, ratio of external surface area to the internal surface area 12.27 19.31 Anff, net flow area per face area 0.534 0.497 Afo, Fin surface area per external surface area 0.839 0.905

Figure 0-19 Correlated external surface heat transfer data for surfaces of table 3-10

From the chosen configuration we can get: 1. The ratio between external surface area and internal surface area (Ao/Ai). 2. The ratio between external surface area and face area per row (Ao/Af Nr). 3. Tube spacing and dimensions. Design procedure: 1. Assume face velocity to be 1.5 m/s. Face area = (Air flow rate / face velocity ). From face velocity and geometry of the coil with air properties at inlet conditions we can use the Colburn J factor to calculate the outside heat transfer coefficient. 2. Assume water inlet and outlet temperatures, Water flow rate = ( Q / (Cp,w (tout tin)). Assuming water velocity inside tubes to be 1 m/s, and knowing the dimensions we may calculate the water side heat transfer coefficient using empirical formulae.

39

Now that we have calculated the inside and outside heat transfer coefficient we can calculate the coil factor. Dividing the coil into some twenty elements we can calculate each element's area assuming each element to have the same capacity. Station ha

(kJ/kg) ti (C) hi

(kJ/kg) Element ha avg hi avg

(kJ/kg) ti avg (C)

Dbt (C)

1 101 ** ** -------- ** ** ** ** 2 1-2 . 20 19-20 Outlet After calculating the outlet air conditions we have to verify that it coincide the required outlet conditions. If not we have to change some of the assumed values like the inlet and outlet water temperatures or water velocity inside tubes and repeat until we get the required outlet conditions.

40

3.5.3 Developed code for coil design {log J_colburn factor = -0.3559192 * log (Re E-3) - 2.06083} J_c = 10^((-0.3559192 * Log10(Re_a/1000)) - 2.06083) {inlet air data = 40 oC dbt with 50 % relative humidity.} {outlet air data = 8.8 oC dbt with 94 % relative humidity. } {air flow rate = 1.085 m3/s} vel_fair = 2 Vdot_air = 1.085 T_o=40 P_=101.325 R_o = 0.5 T_R = 21 R_R = 0.5 T_1 =9.3 R_1 = 1 {assumptions} t_rin = 6 t_rout = 12 v_water = 1 {air properties} v_R=volume(AirH2O,T=T_R ,P=P_,R=R_R ) miu_o=VISCOSITY(AirH2O,T=T_o,P=P_,R=R_o) miu_1 = VISCOSITY(AirH2O,T=T_1,P=P_,R=R_1 ) miu_avg = (miu_o+miu_1)/2 Cp_o = CP(AirH2O,T=T_o,P=P_,R=R_o) Cp_1 = CP(AirH2O,T=T_1,P=P_,R=R_1) Cp_avg = (Cp_o+Cp_1)/2 k_o = CONDUCTIVITY(AirH2O,T=T_o,P=P_,R=R_o) k_1 = CONDUCTIVITY(AirH2O,T=T_1,P=P_,R=R_1 ) k_avg = (k_o+k_1)/2 h_o = ENTHALPY(AirH2O,T=T_o,P=P_,R=R_o) h_1 = ENTHALPY(AirH2O,T=T_1,P=P_,R=R_1) mdot_air =Vdot_air /v_R A_face=Vdot_air/vel_fair Q_cc = mdot_air *(h_o-h_1) {G = mass velocity } G =mdot_air /A_nff/A_face Re_a=D_h*G/miu_avg Pr = (miu_avg*Cp_avg*1000/K_avg) St = J_c/(Pr^(2/3)) {h_c = convection heat transfer coefficient} h_c = St*G*Cp_avg*1000 t_ravg = (t_rin+t_rout)/2 mdot_water = Q_cc /4.18/(t_rout-t_rin) roh_ravg = DENSITY(Water,T=t_ravg,P=200) miu_water = VISCOSITY(Water,T=t_ravg,P=200) Re_w = roh_ravg*v_water*D_i/miu_water Pr_w = PRANDTL(Water,T=t_ravg,P=200) k_f = CONDUCTIVITY(Water,T=t_ravg,P=200) Nu_D = 0.023*(Re_w^(4/5))*(Pr_w^0.4) h_r = Nu_D*K_f/D_i

41

{R_cf = Coil factor } R_cf = h_c*(A_oi)/h_r/Cp_avg y = (t_rout-t_rin)/(h_o-h_1) dpt_o = DEWPOINT(AirH2O,T=T_o,P=P_,R=R_o) hi_dpto =ENTHALPY(AirH2O,T=dpt_o,P=P_,R=1) h_ab=(dpt_o-t_rout+y*h_o+R_cf * hi_dpto)/(R_cf+y) {n = no. of stations x = specific heat transfer through each element.} n=50 x=(h_o-h_1)/(n-1) h_a[1]=h_o DUPLICATE j=2,n h_a[j]=h_a[j-1] - x END t_r[1] = t_rout DUPLICATE j=2,n t_r[j]=t_r[j-1] - (x/4.18/mdot_water *mdot_air) END DUPLICATE j=1,n h_i[j] = 9.3625+1.7861*(t_i[j])+0.01135*(t_i[j])^2+0.00098855*(t_i[j])^3 (t_i[j]/R_cf)-(t_r[j]/R_cf)-h_a[j]+h_i[j]=0 END DUPLICATE j=2,n mdot_air*(h_a[j-1]-h_a[j])=h_c*A_[j]/(Cp_avg*1000)*((h_a[j-1]+h_a[j])/2-(h_i[j-1]+h_i[j])/2) END A_cum[2]=A_[2] DUPLICATE j=3,n A_cum[j] = A_[j]+A_cum[j-1] END A_tot=A_cum[n] dbt_[1] = T_o DUPLICATE j=2,n mdot_air*1000*Cp_avg*(dbt_[j-1]-dbt_[j])=h_c*A_[j]*((dbt_[j-1]+dbt_[j])/2-(t_i[j-1]+t_i[j])/2) END N_r=A_tot/F_s/A_face side_T=A_face^0.5 N_T=side_T/S_T N_Tn=round(N_T) side_Tn=N_Tn*S_T W=A_face/side_Tn side_L=round(N_r)*S_L nooffinperinch = 0.0254/S_f n_circuits=mdot_water/roh_ravg/ v_water/(pi/4*D_i^2) n_c=round(n_circuits) v_watern=mdot_water/roh_ravg/

42

n_c/(pi/4*D_i^2) 3.5.4 Code Output A_face=0.5425 A_nff=0.497 A_oi=19.31 A_tot=112.3 [m2] Cp_1=1.02 [kJ/kg-K] Cp_avg=1.034 Cp_o=1.049 [kJ/kg-K] D_h=0.003865 D_i=0.01588 D_o=0.01717 dpt_o=27.59 [C] F_s=22.86 G=4.77 h_1=27.68 [kJ/kg] h_ab=136.7 h_c=52.13 h_o=100.8 [kJ/kg] h_r=3768 hi_dpto=87.78 [kJ/kg] J_c=0.008692 k_1=0.02437 [W/m-K] k_avg=0.02555 k_f=0.5782 [W/m-K] k_o=0.02673 [W/m-K] mdot_air=1.286

mdot_water=3.75 miu_1=0.00001773 [kg/m-s] miu_avg=0.00001843 miu_o=0.00001913 [kg/m-s] miu_water=0.001345 [kg/m-s] n_c=19 n_circuits=18.95 N_r=9.059 N_T=19.33 N_Tn=19 n=50 nooffinperinch=7.753 Nu_D=103.5 P_=101.3 Pr_w=9.743 Pr=0.746 Q_cc=94.05 R_1=1 R_cf=0.2583 R_o=0.5 R_r=0.5 Re_a=1000

Re_w=11802 roh_ravg=1000 [kg/m^3] S_f=0.003276 S_L=0.04445 S_T=0.0381 side_L=0.4001 side_T=0.7365 side_Tn=0.7239 St=0.01057 T_1=9.3 T_f=0.0004064 T_o=40 T_r=21 t_ravg=9 t_rin=6 t_rout=12 v_R=0.8436 [m^3/kg] v_water=1 v_watern=0.9972 Vdot_air=1.085 vel_fair=2 W=0.7494 x=1.492 y=0.08205

Table 0-11 Air Parameters at each station Station,j ha,j hi,j ti,j tr,j A,j Acum,j Dbt,j

1 100.8 63.41 21.66 12 40

Side_Tn

Side_L W

43

2 99.31 62.46 21.4 11.88 1.026 1.026 39.27 3 97.82 61.51 21.13 11.76 1.041 2.067 38.55 4 96.33 60.56 20.87 11.63 1.057 3.123 37.84 5 94.83 59.63 20.6 11.51 1.073 4.196 37.14 6 93.34 58.69 20.34 11.39 1.09 5.287 36.44 7 91.85 57.77 20.07 11.27 1.108 6.395 35.75 8 90.36 56.84 19.8 11.14 1.127 7.521 35.07 9 88.86 55.93 19.53 11.02 1.146 8.667 34.39

10 87.37 55.02 19.25 10.9 1.166 9.834 33.72 11 85.88 54.11 18.98 10.78 1.188 11.02 33.05 12 84.39 53.21 18.7 10.65 1.21 12.23 32.4 13 82.9 52.32 18.43 10.53 1.233 13.46 31.74 14 81.4 51.43 18.15 10.41 1.258 14.72 31.1 15 79.91 50.55 17.87 10.29 1.283 16.01 30.45 16 78.42 49.67 17.59 10.16 1.31 17.32 29.82 17 76.93 48.8 17.31 10.04 1.339 18.66 29.18 18 75.43 47.93 17.02 9.918 1.369 20.02 28.56 19 73.94 47.07 16.74 9.796 1.401 21.42 27.93 20 72.45 46.21 16.45 9.673 1.434 22.86 27.31 21 70.96 45.36 16.16 9.551 1.469 24.33 26.7 22 69.46 44.52 15.87 9.429 1.507 25.83 26.08 23 67.97 43.68 15.58 9.306 1.547 27.38 25.47 24 66.48 42.85 15.29 9.184 1.589 28.97 24.87 25 64.99 42.02 14.99 9.061 1.634 30.61 24.26 26 63.5 41.2 14.7 8.939 1.683 32.29 23.66 27 62 40.39 14.4 8.816 1.734 34.02 23.06 28 60.51 39.58 14.1 8.694 1.79 35.81 22.47 29 59.02 38.77 13.8 8.571 1.849 37.66 21.87 30 57.53 37.97 13.5 8.449 1.913 39.58 21.28 31 56.03 37.18 13.2 8.327 1.983 41.56 20.68 32 54.54 36.39 12.89 8.204 2.058 43.62 20.09 33 53.05 35.61 12.59 8.082 2.14 45.76 19.5 34 51.56 34.83 12.28 7.959 2.229 47.99 18.91 35 50.07 34.06 11.97 7.837 2.327 50.31 18.32 36 48.57 33.3 11.66 7.714 2.435 52.75 17.72 37 47.08 32.54 11.35 7.592 2.554 55.3 17.13 38 45.59 31.78 11.03 7.469 2.686 57.99 16.54 39 44.1 31.03 10.72 7.347 2.835 60.82 15.94 40 42.6 30.29 10.4 7.224 3.001 63.82 15.34 41 41.11 29.55 10.09 7.102 3.19 67.01 14.74 42 39.62 28.82 9.769 6.98 3.406 70.42 14.14 43 38.13 28.09 9.45 6.857 3.654 74.07 13.54 44 36.63 27.37 9.129 6.735 3.944 78.02 12.93 45 35.14 26.65 8.806 6.612 4.287 82.3 12.31 46 33.65 25.93 8.483 6.49 4.698 87 11.69 47 32.16 25.23 8.158 6.367 5.199 92.2 11.07 48 30.67 24.52 7.832 6.245 5.824 98.03 10.44 49 29.17 23.82 7.504 6.122 6.626 104.7 9.803 50 27.68 23.13 7.176 6 7.691 112.3 9.158

44

3.5.5 The effect of different variables 3.5.5.1 Effect of face velocity In this study, the face velocity is being changed from 0.5 to 3 m / s and the effect on the outside total heat transfer area, and number of rows are plotted. Table 0-12 Face Velocity Effect on Coil dimensions

Run Face

velocity (m/s)

h_c (W/m2K)

Outlet dry bulb

(C)

Total outside

area (m2)

No. of

rows

side_L (m)

Side_tn (m)

Width (m)

1 0.5 21.35 9.127 211.6 4.266 0.1778 1.486 1.46 2 0.6 24.01 9.132 192.9 4.666 0.2223 1.334 1.356 3 0.7 26.51 9.135 178.7 5.043 0.2223 1.257 1.233 4 0.8 28.89 9.138 167.5 5.402 0.2223 1.181 1.148 5 0.9 31.17 9.14 158.4 5.747 0.2667 1.105 1.091 6 1 33.36 9.142 150.8 6.081 0.2667 1.029 1.055 7 1.1 35.47 9.144 144.4 6.405 0.2667 0.9906 0.9957 8 1.2 37.52 9.146 138.9 6.722 0.3112 0.9525 0.9493 9 1.3 39.5 9.148 134.1 7.031 0.3112 0.9144 0.9127

10 1.4 41.43 9.15 129.9 7.334 0.3112 0.8763 0.8844 11 1.5 43.32 9.151 126.2 7.632 0.3556 0.8382 0.863 12 1.6 45.16 9.153 122.9 7.925 0.3556 0.8382 0.809 13 1.7 46.95 9.154 119.8 8.214 0.3556 0.8001 0.7977 14 1.8 48.71 9.155 117.1 8.499 0.3556 0.762 0.791 15 1.9 50.44 9.157 114.6 8.781 0.4001 0.762 0.7494 16 2 52.13 9.158 112.3 9.059 0.4001 0.7239 0.7494 17 2.1 53.8 9.159 110.2 9.334 0.4001 0.7239 0.7137 18 2.2 55.44 9.16 108.3 9.607 0.4445 0.6858 0.7191 19 2.3 57.05 9.162 106.5 9.877 0.4445 0.6858 0.6879 20 2.4 58.63 9.163 104.8 10.14 0.4445 0.6858 0.6592 21 2.5 60.19 9.164 103.3 10.41 0.4445 0.6477 0.6701 22 2.6 61.73 9.165 101.8 10.67 0.489 0.6477 0.6443 23 2.7 63.25 9.166 100.5 10.94 0.489 0.6477 0.6204 24 2.8 64.75 9.168 99.18 11.2 0.489 0.6096 0.6357 25 2.9 66.23 9.169 97.97 11.46 0.489 0.6096 0.6137 26 3 67.69 9.17 96.83 11.71 0.5334 0.6096 0.5933

45

Figure 0-20 The Required Coil External Surface Area and No. of Rows Variation with Face Velocity When making the choice of the face velocity, the number of rows should be considered as well as the total surface area, a smaller area with larger number of rows is not a favorable case as the coil would be considered bulky and would thus increase both the air side pressure drop as well as the water side pressure drop. And thus the initial cost would be small (smaller surface area and therefore lower material weight), but the running cost would be larger (higher pressure drop). Finally a compromised solution should be reached based on economical considerations. From the plot shown in figure 3-20 it can be easily concluded that as the face velocity increases, the coil face area decreases but the coil depth increase but the total outside surface area is being reduced. From the graph shown in figure 3-20, the chosen air velocity is such that the rate of total external area change with face velocity is small. That is why we tried to take the face velocity around 2.2m/s, but this value results in a number of rows of 9.6. this value couldn't be achieved that is why the choice was for the 2 m/s face velocity which results in a number of rows of 9.

1 1.4 1.8 2.2 2.6 390

100

110

120

130

140

150

160

6

7

8

9

10

11

12

velfair

Ato

t [m

2]

Nr

46

3.5.5.2 Effect of water velocity In this study, the water velocity is being changed from 0.5 to 1.5 m / s while the face velocity is kept constant at 2 m/s. Table 0-13 Water Velocity Effect on Coil Dimensions and No. of Circuits

Run Water

velocity (m/s)

h_r (W/m2K)

Outlet dry bulb (C)

Total outside

area (m2)

No. of rows

side_L (m)

No. of circuits

1 0.5 2164 9.187 145.9 11.76 0.5334 38 2 0.54 2302 9.183 141.1 11.38 0.489 35 3 0.58 2437 9.18 137 11.04 0.489 33 4 0.62 2571 9.176 133.3 10.75 0.489 31 5 0.66 2703 9.174 130.1 10.49 0.4445 29 6 0.7 2833 9.171 127.2 10.25 0.4445 27 7 0.74 2962 9.169 124.6 10.04 0.4445 26 8 0.78 3089 9.167 122.2 9.853 0.4445 24 9 0.82 3215 9.165 120 9.679 0.4445 23

10 0.86 3340 9.163 118.1 9.521 0.4445 22 11 0.9 3464 9.161 116.3 9.375 0.4001 21 12 0.94 3586 9.16 114.6 9.241 0.4001 20 13 0.98 3708 9.159 113.1 9.117 0.4001 19 14 1.02 3829 9.157 111.6 9.002 0.4001 19 15 1.06 3948 9.156 110.3 8.896 0.4001 18 16 1.1 4067 9.155 109.1 8.796 0.4001 17 17 1.14 4185 9.154 107.9 8.703 0.4001 17 18 1.18 4302 9.153 106.8 8.615 0.4001 16 19 1.22 4418 9.152 105.8 8.533 0.4001 16 20 1.26 4534 9.151 104.9 8.456 0.3556 15 21 1.3 4649 9.15 104 8.383 0.3556 15 22 1.34 4763 9.149 103.1 8.315 0.3556 14 23 1.38 4876 9.149 102.3 8.25 0.3556 14 24 1.42 4989 9.148 101.5 8.188 0.3556 13 25 1.46 5101 9.147 100.8 8.129 0.3556 13 26 1.5 5212 9.146 100.1 8.074 0.3556 13

47

3.5.5.3 Effect of inlet water temperature In this study, the inlet water temperature is being changed from 5 to 8 C while the face and the water velocities is kept constant at 2 and 1 m/s respectively. Table 0-14 Inlet Water Temperature Effect on Coi l Dimensions

Run t_rout C t_rin C

h_r (W/m2K)

Outlet dry bulb (C)

Total outside

area (m2)

No. of rows

side_L (m)

No. of circuits

1 12 5 3743 9.113 103 8.305 0.3556 16 2 12 5.12 3746 9.119 104 8.384 0.3556 17 3 12 5.24 3749 9.124 105 8.466 0.3556 17 4 12 5.36 3752 9.13 106 8.55 0.4001 17 5 12 5.48 3755 9.135 107.1 8.638 0.4001 17 6 12 5.6 3758 9.14 108.3 8.729 0.4001 18 7 12 5.72 3761 9.146 109.4 8.823 0.4001 18 8 12 5.84 3764 9.151 110.6 8.922 0.4001 18 9 12 5.96 3767 9.156 111.9 9.024 0.4001 19

10 12 6.08 3770 9.161 113.2 9.13 0.4001 19 11 12 6.2 3773 9.167 114.6 9.241 0.4001 20 12 12 6.32 3777 9.172 116 9.358 0.4001 20 13 12 6.44 3780 9.177 117.6 9.479 0.4001 20 14 12 6.56 3783 9.182 119.1 9.607 0.4445 21 15 12 6.68 3786 9.187 120.8 9.741 0.4445 21 16 12 6.8 3789 9.192 122.6 9.882 0.4445 22 17 12 6.92 3792 9.197 124.4 10.03 0.4445 22 18 12 7.04 3795 9.201 126.4 10.19 0.4445 23 19 12 7.16 3798 9.206 128.4 10.36 0.4445 23 20 12 7.28 3801 9.211 130.6 10.53 0.489 24 21 12 7.4 3804 9.216 133 10.72 0.489 25 22 12 7.52 3807 9.22 135.5 10.93 0.489 25 23 12 7.64 3810 9.225 138.2 11.15 0.489 26 24 12 7.76 3813 9.229 141.2 11.38 0.489 27 25 12 7.88 3816 9.233 144.4 11.64 0.5334 28 26 12 8 3819 9.238 147.9 11.92 0.5334 28

48

3.5.5.4 Effect of ambient air temperature In this study, all the coil design parameters are being constant except for the outside air dry-bulb temperature. The results are shown in table 3-15 for an outside temperature variation from 21 C to 44 C. Table 0-15 Oustide air condition effect on coil capacity

run To Outside RH Atot Qcc Nr Exit dry bulb 1 21 0.5 75.71 16.83 6.105 9.541 3 22 0.5 76.79 19.76 6.192 9.525 5 23 0.5 78.43 22.79 6.324 9.502 7 24 0.5 80.32 25.91 6.477 9.476 9 25 0.5 82.33 29.13 6.638 9.45 11 26 0.5 84.38 32.47 6.804 9.424 13 27 0.5 86.46 35.91 6.972 9.399 15 28 0.5 88.53 39.48 7.139 9.375 17 29 0.5 90.6 43.17 7.305 9.352 19 30 0.5 92.64 46.99 7.47 9.33 21 31 0.5 94.68 50.96 7.634 9.309 23 32 0.5 96.69 55.06 7.797 9.289 25 33 0.5 98.69 59.32 7.958 9.269 27 34 0.5 100.7 63.74 8.117 9.251 29 35 0.5 102.6 68.32 8.276 9.234 31 36 0.5 104.6 73.08 8.434 9.217 33 37 0.5 106.5 78.02 8.591 9.201 35 38 0.5 108.5 83.16 8.747 9.186 37 39 0.5 110.4 88.5 8.903 9.172 39 40 0.5 112.3 94.05 9.059 9.158 41 41 0.5 114.3 99.82 9.214 9.145 43 42 0.5 116.2 105.8 9.37 9.132 45 43 0.5 118.1 112.1 9.525 9.12 47 44 0.5 120.1 118.6 9.681 9.109

Weather Condition effect on Coil Capacity and No. of Rows

0

20

40

60

80

100

120

140

21 23.5 26 28.5 31 33.5 36 38.5 41 43.5

Outside Dry-Bulb Temperature (deg.C)

Coil

Capa

city

(kW

)

6

6.5

7

7.5

8

8.5

9

9.5

10

No. O

f Row

s

CoilCapacityNo. ofrows

49

50

3.6 DDC Control System For Total Fresh Air AHU 3.6.1 Sequence of Operation The AHU consists of filter section including pre and bag filters, cooling coil, electric heater, steam humidifier and supply fan with variable speed drive. 1. Start of the supply fan according to a programmable time program with the possibility of exception programs for holidays, maintenance, etc.. 2. Supply fan start interlocks fresh air damper to open to preset location. 3. supply fan flow, indicated by differential pressure switch interlocks control system start so that control functions are not performed if fan is not in normal operation. Flow failure will initiate an alarm at the control system. 4. Temperature and humidity is measured by sensor mounted on supply air duct. 5. Temperature and humidity are then compared to the previously adjusted set points & controller signals 3-way cooling valveto modulate for cooling or dehumidification, or current valve mounted on electric heater to modulate for heating or reheat, and steam humidifier to modulate for humidification upon deviation from the adjusted set point. 6. Overheat thermostat is mounted downstream of the electric heater. An alarm will be issued in case of overheat sensed. 7. Status , and trip alarm of supply fan will be monitored through control system. Failure to give a status after a start signal was issued will signal an alarm after a dedicated time delay. 8. Differential pressure switch mounted across filterbanks will issue an alarm in case of filter dirty. 9. Pressure sensor mounted on supply air duct measures supply air pressure, in case of pressure drop due to absolute filter being dirty, variable speed drive will be signaled to increase the fan motor speed to ensure a constant pressure of air flow to the controlled zone. This is to avoid cross contamination between the controlled zone and surrounding zones due to infiltration, doors opening etc.. 10. Smoke detector mounted on supply air duct will issue an alarm in case of smoke sensed. Control system will be signaled to stop immediately. In case of