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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
3
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
25
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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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22
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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23
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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24
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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25
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
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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
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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.
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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.
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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.
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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
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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,
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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
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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)
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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