Memoria Investigaciones en Ingeniería, núm. 26 (2024). pp. 2-37
https://doi.org/10.36561/ING.26.2
ISSN 2301-1092 • ISSN (en línea) 2301-1106 – Universidad de Montevideo, Uruguay
Este es un artículo de acceso abierto distribuido bajo los términos de una licencia de uso y distribución CC BY 4.0. Para ver una
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Memoria Investigaciones en Ingeniería, núm. 26 (2024). pp. 2-37
https://doi.org/10.36561/ING.26.2
ISSN 2301-1092 • ISSN (en línea) 2301-1106 – Universidad de Montevideo, Uruguay
Este es un artículo de acceso abierto distribuido bajo los términos de una licencia de uso y distribución CC BY 4.0.
Para ver una copia de esta licencia visite https://creativecommons.org/licenses/by/4.0/
Revitalizing Comfort: Designing an Energy–Efficient HVAC System for the
University Auditorium
Comodidad revitalizante: Diseño de un Sistema HVAC Energéticamente Eficiente
para el Auditorio Universitario
Revitalizando o Conforto: Projeto de um Sistema HVAC com Eficiência Energética
para o Auditório Universitário
Abdul Samad Khan
1
(*), Muhammad Ehtesham ul Haque
2
, Adeel Ahmed Khan
3
,
Syed Izhar ul haque
4
, Syed Obaidullah
5
, Muhammad Umer Khan
6
Recibido: 29/07/2023 Aceptado: 13/04/2024
Summary. - Nowadays, thermal comfort is becoming a major problem for people due to increasing global warming
and climatic changes, but it can be resolved by the concept of Heating, Ventilation, and Air Conditioning (HVAC)
systems. The purpose of HVAC is to provide occupants with a comfort zone so that they can feel comfortable according
to their thermal comfort. The core objective of this study is to design and propose an HVAC system as per actual
design conditions for the University Auditorium located in Karachi, Pakistan. A direct Expansion (DX – Type) system
is installed in the Auditorium that has exceeded the lifespan of twenty years, refrigerant R-22 which is currently being
used has been obsolete due to its high GWP (Global Warming Potential) and ODP (Ozone Depletion Potential) values
which are 1810 and 0.05 respectively. To achieve the objective of this study, two approaches are employed. Cooling
Load Temperature Difference (CLTD) method & Hourly Analysis Program (HAP) software. The cooling load
calculated from the CLTD method is 202 kW equivalent to 57.5 Ton of Refrigeration (TR). On the other side, the
cooling load calculated from HAP software is 192.8 kW equivalent to 55 TR. By considering the calculated cooling
load for the University Auditorium, two different HVAC systems are proposed, based on Water cooled and Air-cooled
Vapor Compression Cycle. After this study, engineers will be able to design an HVAC system for any facility as per
design conditions. Also, they can propose different cost-effective and energy-efficient HVAC systems for that
particular space.
Keywords: HVAC; Auditorium; Duct sizing; Cooling load; Piping.
(*) Corresponding Author
1
Lecturer, Department of Mechanical Engineering, NED University of Engineering and Technology (Pakistan), abdulsamadkhan@neduet.edu.pk,
ORCID iD: https://orcid.org/0009-0005-5449-635X
2
Assistant Professor. Department of Mechanical Engineering, NED University of Engineering and Technology (Pakistan),,
mehaque@neduet.edu.pk, ORCID iD: https://orcid.org/0000-0001-8751-348X
3
Assistant Professor. Department of Mechanical Engineering, NED University of Engineering and Technology (Pakistan),
adeelahmedk@neduet.edu.pk, ORCID iD: https://orcid.org/0009-0004-6790-8176
4
Senior Undergraduate Student. Department of Mechanical Engineering, NED University of Engineering and Technology (Pakistan),
izharmumshad97@hotmail.com, ORCID iD: https://orcid.org/0009-0002-5693-5879
5
Senior Undergraduate Student. Department of Mechanical Engineering, NED University of Engineering and Technology (Pakistan),
syedobaid2121@gmail.com, ORCID iD: https://orcid.org/0009-0000-0168-6196
6
Senior Undergraduate Student. Department of Mechanical Engineering, NED University of Engineering and Technology (Pakistan),
engineerumerkhan@gmail.com, ORCID iD: https://orcid.org/0009-0001-7173-6395
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Resumen. - Hoy en día, el confort térmico se está convirtiendo en un gran problema para las personas debido al
aumento del calentamiento global y los cambios climáticos, pero puede ser resuelto por el concepto de sistemas de
Calefacción, Ventilación y Aire Acondicionado (HVAC). El objetivo de HVAC es proporcionar a los ocupantes una
zona de confort para que puedan sentirse cómodos de acuerdo con su confort térmico. El objetivo central de este
estudio es diseñar y proponer un sistema HVAC según las condiciones de diseño reales para el Auditorio Universitario
ubicado en Karachi, Pakistán. En el Auditorio se encuentra instalado un sistema de Expansión Directa (Tipo DX) que
ha superado la vida útil de veinte años, el refrigerante R-22 que se utiliza actualmente ha quedado obsoleto por su
alto GWP (Global Warming Potential) y ODP (Ozone Depletion Potencial) valores que son 1810 y 0.05
respectivamente. Para lograr el objetivo de este estudio, se emplean dos enfoques. Método de diferencia de
temperatura de carga de enfriamiento (CLTD) y software de programa de análisis por hora (HAP). La carga de
refrigeración calculada a partir del método CLTD es de 202 kW equivalente a 57,5 Toneladas de Refrigeración (TR).
Por otro lado, la carga de refrigeración calculada a partir del software HAP es de 192,8 kW equivalente a 55 TR. Al
considerar la carga de enfriamiento calculada para el Auditorio Universitario, se proponen dos sistemas HVAC
diferentes, basados en el ciclo de compresión de vapor enfriado por agua y enfriado por aire. Después de este estudio,
los ingenieros podrán diseñar un sistema HVAC para cualquier instalación según las condiciones de diseño. Además,
pueden proponer diferentes sistemas HVAC rentables y energéticamente eficientes para ese espacio en particular.
Palabras clave: HVAC; Auditorio; Dimensionamiento de ductos; Carga de enfriamiento; Piping.
Resumo. - Hoje em dia, o conforto térmico está a tornar-se um grande problema para as pessoas devido ao aumento
do aquecimento global e às alterações climáticas, mas pode ser resolvido pelo conceito de sistemas de Aquecimento,
Ventilação e Ar Condicionado (HVAC). O objetivo do HVAC é proporcionar aos ocupantes uma zona de conforto
para que se sintam confortáveis de acordo com o seu conforto térmico. O objetivo principal deste estudo é projetar e
propor um sistema HVAC de acordo com as condições reais de projeto para o Auditório Universitário localizado em
Karachi, Paquistão. No Auditório está instalado um sistema de Expansão Direta (Tipo DX) que ultrapassou sua vida
útil de vinte anos. O refrigerante R-22 atualmente utilizado tornou-se obsoleto devido ao seu alto GWP (Potencial de
Aquecimento Global) e ODP (Depleção de Ozônio). Potencial) valores que são 1810 e 0,05 respectivamente. Para
atingir o objetivo deste estudo, duas abordagens são utilizadas. Método de diferença de temperatura de carga de
resfriamento (CLTD) e software de programa de análise horária (HAP). A carga de resfriamento calculada a partir
do método CLTD é de 202 kW equivalente a 57,5 toneladas de refrigeração (TR). Por outro lado, a carga de
refrigeração calculada a partir do software HAP é de 192,8 kW equivalente a 55 TR. Ao considerar a carga de
refrigeração calculada para o Auditório Universitário, são propostos dois sistemas HVAC diferentes, baseados no
ciclo de compressão de vapor refrigerado a água e arrefecido a ar. Após este estudo, os engenheiros serão capazes
de projetar um sistema HVAC para qualquer instalação com base nas condições de projeto. Além disso, eles podem
propor diferentes sistemas HVAC econômicos e energeticamente eficientes para esse espaço específico.
Palavras-chave: AVAC; Público; Dimensionamento de dutos; Carga de resfriamento; Tubulação.
Nomenclature:
A Area
ACH Air Changes per Hour
Cooling Load Temperature Difference Adjusted
D Diameter
E Efficiency
Ballast Factor
Utilization Factor
g Acceleration due to gravity
Inside Relative Humidity
Head Loss
Mixed air enthalpy
Outside air enthalpy
Recirculated air enthalpy
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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l Length
Equivalent Length
Mixed air flowrate
Outdoor air flowrate
Recirculated air flowrate
N No. of Occupant
Heat Gain
Latent Heat Gain by Infiltration
Sensible Heat Gain by Infiltration
Latent Heat Gain
Sensible Heat Gain
Latent Heat Gain by Ventilation
Sensible Heat Gain by Ventilation
Temperature below the floor of the Auditorium
Outside average temperature
Inside design temperature
Mixed air temperature
Outside design temperature
Supply air temperature
R Thermal Resistance
Area Outdoor Air rate
People's Outdoor Air rate
U Thermal Transmittance
V Volume
v Velocity
Infiltration Air Flowrate
Minimum Outdoor Air Flowrate
Outside Air Flowrate
Recirculation Air Flowrate
Ventilation Air Flowrate
W Wattage
Inside Humidity Ratio
Outside Humidity Ratio
Greek Symbols:
ρ Density
Pressure drop
Relative Humidity
Heat gain through wall or roof, at calculation hour
Time
Time interval
Sol-air temperature at time
Constant indoor room temperature
Conduction transfer function coefficients
Subscripts:
a adjacent space
adj adjusted
average
inside
il infiltration latent
is infiltration sensible
M mixed
m minimum outdoor air
n summation index (each summation has as many terms as there are non-negligible values
of coefficients)
outside
P people
R recirculated
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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r recirculation
S supply
vl ventilation latent
vs ventilation sensible
Acronyms
AHU Air Handling Unit
ASHRAE American Society of Heating, Refrigeration & Air Conditioning Engineers
CLF Cooling Load Factor
CLTD Cooling Load Temperature Difference
DX Direct Expansion
EER Energy Efficiency Ratio
GWP Global Warming Potential
HAP Hourly Analysis Program
HFC Hydrofluorocarbons
HVAC Heating Ventilation and Air Conditioning
IAQ Indoor Air Quality
LHG Latent Heat Gain
ODP Ozone Depletion Potential
PET Polyethylene terephthalate
SCL Solar Cooling Load Factor
SEER Seasonal Energy Efficiency Ratio
SHG Sensible Heat Gain
TFM Transfer Function Method
TR Ton of Refrigeration
VCC Vapor Compression Cycle
1. Introduction. - Nowadays, global warming become one of the major issues. If temperature, pressure, and relative
humidity in the ambient atmospheric conditions are observed, there is an uncomfortable situation for the people. There
are many regions in the World where outside ambient conditions are too hot and humid. So, in this situation, people
can’t perform daily life activities and tasks due to their lower thermal comfort. Due to the environmental changes, the
term “climate” comes into play [1]. Pakistan, the country in this World is now tolerating the summer season extending
from April to November. At the global scale, it is significantly observed there was a greater number of hot days as
compared to cold days over the past decade which is evident the frequency of hot and humid days is higher. Also, from
the study conducted by the Pakistan Meteorological Department, a significant increase in heatwave days is observed
which is now a major issue for Pakistan especially the occupants living in different cities due to thermal comfort zones
[2].
Without people's thermal comfort, the occupants can’t feel comfortable, and they can’t perform daily activities. By
considering the mentioned situation and conditions related to human comfort, especially in hot climatic conditions,
scientists developed the concept of an HVAC system which is the most important requirement for people.
The concept of thermal comfort is a state of mind, the essential parameter for the occupant’s comfort zone that provides
satisfaction to the occupant so that one can perform his tasks in a comfortable environment [3-5]. The comfort zone
for an occupant is predicted by relative humidity, air velocity, air & radiant temperatures, clothing insulation, and
metabolic rate [6], it can be defined by a range of operative temperatures that will provide acceptable thermal
conditions for a person's state of mind [7].
An HVAC system is mainly responsible for maintaining the desired IAQ by supplying adequate and acceptable fresh
air [8, 9]. HVAC systems need to be much more efficient as it consumes around 60% of the building’s total energy
consumption [10]. In Pakistan, it is observed that the systems which are installed for human comfort are not designed
on standard conditions. Either the system is overdesigned in that it produces too much cooling effect and is not
economically feasible or the system is under designed so that the occupants are not thermally comfortable [11].
Heat gain is the heat generated by material or equipment in space. HVAC system performance depends on the heat
generated by several pieces of equipment. Heat gain ultimately depends on several factors such as room orientation
relative to solar radiation, electric devices or appliances, wall and roof materials, and the number of occupants present
in a space [12].
Nowadays, due to increasing heat-generating sources and hot & humid climatic conditions, the need for an air
conditioning system is a must. On a domestic level, air conditioning requirement is fulfilled by split air conditioners
[13] but in large buildings, offices, or auditoriums, these are not feasible due to insufficient supply of air flowrate. So,
to resolve this issue, the HVAC system plays an important role to supply adequate fresh air and maintain IAQ within
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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the facility. Researchers and designers introduced some software to make energy-efficient systems. But the problem
with this software is it doesn’t completely fit the real-time data and the real-time data is something else that is different
from the built-in values of the software. This is the reason why when researchers design a particular system that is only
based on software and no manual method is used, there must be an error [14].
There are several ways to determine the cooling load of space. Some methods such as CLTD, CLF, SCL, and HAP
(Transfer Function) are being used for cooling load calculation. Each of these methods possesses a different and unique
methodological approach for calculating space cooling load. CLTD and HAP methods will be discussed in this paper
[15].
CLTD is widely used for manual cooling load calculation, as proposed by ASHRAE. This method is used to calculate
space cooling loads in which heat-dissipating devices are present. It is also used to calculate the load due to heat
dissipation from the walls, windows, and roofs in the space [16]. For walls and roofs, CLTD uses the concept of heat
transfer temperature difference but for internal load and windows, it uses CLF [17]. The tabulated CLTD and CLF data
were calculated using the transfer function method, which yielded cooling loads for standard environmental conditions
and zone types. The cooling loads for each component are then summed to obtain the total zone cooling load [18]. The
cooling load of an auditorium space is determined using the CLTD/CLF method, developed as a manual calculation
technique relying on tabulated CLTD and CLF values. These tabulated data were derived using the transfer function
method, providing cooling load estimates for typical environmental conditions and zone types. These loads were
subsequently standardized for easy hourly calculation by designers through normalization. Total zone cooling load is
computed by summing the cooling loads of individual components.
HAP is a powerful computer-based tool developed by Carrier Corporation, designed for consulting engineers, HVAC
contractors, facility engineers, and other professionals involved in the design and analysis of commercial building
HVAC systems. The HAP uses the ASHRAE Transfer Function Method (TFM) for system load calculations and
detailed 8,760 hour-by-hour simulation techniques for energy analysis. The TFM uses transfer functions to model
transient heat transfer equations [19].
The two major functions of HAP are:
1. Estimating load and designing systems
2. Performing Energy and Cost analysis
The HAP software uses a specific built-in program called the “HAP System Design Load” program to calculate, design,
and size the HVAC system. The following are some features of this program:
Calculates design cooling and heating loads for spaces, zones, and coils in the HVAC system.
Determines required airflow rates for spaces, zones, and the system.
Sizes cooling and heating coils.
Sizes air circulation fans.
Sizes chillers and boilers.
Some previous case studies were performed by different researchers for cooling load calculation of particular facilities
by different methods and there is a significant result variation after validation by researchers which leads to an
inappropriate design of the HVAC system. ALameen awad Alameen et al. [20] performed a study related to the
designing of an Air Conditioning System for a Sports Hall, with a capacity of 1000 occupants in Sudan by using CLTD
and HAP methods. The results obtained from CLTD and HAP were 116 TR and 103 TR respectively, with a percentage
error of 13%. Khakre et al. [21] conducted a study for cooling load calculation that incorporates the CLTD method for
an evaporative cooling system. The cooling load obtained from the CLTD method was 42.35 TR and from the HAP
program, it was 38.6 TR, contributing an error of 9% which is not allowed to design an accurate HVAC system. These
variations of calculated cooling load from two different methods make it impossible to design an accurate HVAC
system.
Based on cooling load calculation, all previous research studies are focused on designing of HVAC system for any
facility by using the CLTD method and HAP software. The purpose of this research study is to design and propose
different possible HVAC systems for the University Auditorium. A Central Air Conditioning system is already
installed in the Auditorium which is DX – type system. An HVAC system that is currently being operated has exceeded
the lifespan of twenty years and Refrigerant R-22 (Chloro-Difluoro-Methane) is currently being used and has been
obsolete due to high GWP and ODP values. In the existing HVAC system, R-22 enters AHU and cools the air without
the need for secondary refrigerant but it’s a danger zone for occupants if the refrigerant piping leaks for some reason,
it will directly be mixed with cool air and goes into the space where potential health hazards for occupants will be
taken place. So, R-22 (Freon) is not suitable for the cause rather it will be harmful to the occupants. To the knowledge
of the authors, no study is available that proposed new HVAC systems with the refrigerant R-410 (A) for the University
Auditorium. R-410 (A) is a zeotropic and its ODP is zero due to the absence of chlorine. Although it has a higher GWP
than R-22 due to its higher SEER rating, and by reducing the power consumption of the system, it is overall more
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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environmentally friendlier than R-22. The novelty of this research work is to fill this research gap. In this paper, the
cooling load of the University Auditorium is calculated via the CLTD method and HAP software. The calculated
cooling load is then used to select new equipment like Chillers for both proposed systems which are water-cooled and
air-cooled. Other equipment like AHU, cooling towers, and pumps are also selected for both systems as per available
data and calculations.
2. The University Auditorium. - The analysis involves the designing of an HVAC system for the University
Auditorium located in Karachi. The Auditorium contains electrical equipment and devices that are dissipating
continuous heat for which proper temperature control is necessary and IAQ must be maintained. The audience or
working staff can’t feel comfortable in such an environment when there is no proper control of inside dry bulb
temperature and relative humidity due to occupants present at that particular time. DX – Type system has already been
installed in the Auditorium and this system provides both air-conditioning as well as ventilation through a ducting
network in which refrigerant directly enters AHU and cools the air without the need for any secondary refrigerant. This
system has exceeded the lifespan of twenty years with the usage of R-22 which is now obsolete due to its high GWP
and ODP values which are 1810 and 0.05 respectively. Montreal and Kyoto Protocols introduced HFCs for the
replacement of chlorofluorocarbons (CFCs) and hydrochlorofluorocarbons (HCFCs) because of the low ODP of
HCFCs [22]. However, the qualifying criteria for the selection of refrigerant is not only based on ODP but the potential
alternative refrigerant has been selected based on a high energetic performance and GWP [23, 24]. Table I represents
the specification of components installed in the HVAC system of the University Auditorium.
Table I. Existing HVAC system of the University Auditorium
The occupancy schedule is taken as 6 hours (from 0800 hours to peak time of 1400 hours). The geographical location
of the Auditorium is as: Latitude is 24.9 °N, Longitude is 67.1 °E and Elevation is 51 meters. The orientation of all
four walls of the Auditorium in a cardinal direction is shown in Figure I.
Components
Specifications
Compressor
Company: Carlyle Carrier
Type: Open-Drive Reciprocating
Model: 5H66
Condenser
Shell & Tube Type
Evaporator
DX - Type
Expansion device
Thermal Expansion Valve
Air Handling Unit
Size: 400 x 300 x 175 cm
Coil Area: 300 x 170 cm
Cooling Tower
Induced draft Counter Flow
Refrigeration cycle
Vapor Compression
Refrigerant
R-22 (Chloro-Difluoro Methane)
Other capacities
Compressor: 75 hp
AHU fan: 15 hp
Cooling tower pump: 1.5 hp
Condenser water pump: 9.8 hp
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Figure I.- Walls orientation of the University Auditorium
The Auditorium consists of a Hall with roof and floor areas are 615.25 m2 & 615.24 m2 respectively. It consists of four
walls South East (Back wall), South West (Sidewall), North East (Sidewall), and North West (Stage wall) with areas
of 67.23 m2, 168.84 m2, 168.84 m2 & 103.79 m2 respectively. Although there is no window in the Auditorium, its
cooling load won’t be calculated. The outside weather data condition for the University Auditorium is obtained by
keeping Jinnah International Airport as a reference [25]. The weather data shows an outside design temperature of 38.9
℃ and an average humidity ratio of 19.8 g/kgda [25]. Figure II depicts how the existing HVAC system of the
Auditorium works while Figure III shows the 3D model drawn on Autodesk Revit.
Figure II.- Process flow diagram of the existing HVAC system of the University Auditorium
(a) (b)
Figure III.- 3D Revit model of the University Auditorium
Figures IV, V, VI & VII represent the front view, back view, top view, and section view of the Auditorium.
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Figure IV.- Front side view of the University Auditorium
Figure V.- Back view of the University Auditorium
Figure VI.- Top view of the University Auditorium
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Figure VII.- Section view of the University Auditorium
3. Methodology. - This section consists of the calculations of the Roof, walls, and floor areas of the Auditorium
manually. Inside weather conditions of the University Auditorium will be selected as per design conditions [26]. Figure
VIII represents the methodological flow chart.
Figure VIII.- Methodological approach
3.1 Manual Cooling Load Estimation (CLTD Method). - Table II depicts the important design conditions for cooling
load calculation.
Inside Condition
Symbols
Values
References
Dry bulb temperature
Relative Humidity
Humidity Ratio
22 °C
60 %
9.9 g/kgda
[26]
[26]
[26]
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Table II. Inside and outside weather conditions
Table III represents surface film coefficients/resistances [25]. Table IV represents the thermal resistances of
Auditorium structural materials.
Table III. Thermal Resistances of Surface Films
Table IV. Thermal Resistances of Structure Materials
3.1.1 Roof. - The cooling load from the roof can be calculated by eq (1) [28].
(1)
U-value for the surface of the roof can be calculated by eq (2). Figure IX shows the thermal network diagram for the
construction material of the roof.
(2)
Outside Condition
Symbols
Values
References
Dry bulb temperature
Average temperature
Humidity Ratio
38.9 °C
36 °C
19.8 g/kgda
[25]
[25]
[25]
Position of surface
The direction of heat
flow
Thermal Resistance at
Emissivity (90%) (m2
k/W)
References
Indoor Vertical
Horizontal
Horizontal
Horizontal
Downward
Upward
0.12
0.16
0.11
[25]
[25]
[25]
Outdoor
(any position for summer at 3.4 m/s)
Any
0.044
[25]
Material
Thermal
Conductivity
(W/ m.k)
Thickness of
material
(L)
Thermal
Resistance
(m2 k/W)
Ref
The air between the roof and ceiling
26.24
60’’
0.058
[27]
The air between the side walls
26.24
54’’
0.052
[27]
Plaster
0.72
1’’
0.035
[25]
Concrete Slab
1.818
6’’
0.0825
[28]
Acoustic Tiles
0.052
5/8’’
0.288
[25]
Hardwood
0.18
1’’
0.141
[25]
Hollow Block
0.88
6’’
0.172
[29]
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Figure IX.- Thermal network diagram for the roof’s construction
Table V depicts the thermal resistance values of the material used in the construction of the Roof.
Table V. Thermal Resistances for Roof Material
The total area of the roof is 615.25 m2. Solar time of 1400 hours with Type-4 roof with suspended ceiling is selected.
The CLTD value obtained is 16 ℃ [28]. value is determined from eq (3) [28].
(3)
3.1.2 Walls. - The cooling load from the wall can be calculated from eq (1) [28]. U-value can be calculated for the
surface of the wall by eq (2). Figures X, XI, XII, and XIII illustrate thermal network diagrams for walls oriented in the
southeast, southwest, northeast, and northwest directions, respectively.
3.1.2.1 South East (Back Wall). - Table VI represents the thermal resistance values of the material used in the
construction of South East (Back Wall).
Figure X.- Thermal network diagram for South East wall construction
Resistance
Description
Value (m2 k/W)
References
Outside Air Resistance
0.044
[25]
Plaster
0.035
[25]
Concrete Slab
0.083
[28]
Air Gap
0.058
[27]
Acoustic tile
0.288
[25]
Inside Air Resistance
0.160
[25]
Resistance
Description
Value (m2 k/W)
References
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Table VI. Thermal Resistances for South East material
3.1.2.2 South West (Side Wall). - Table VII represents the thermal resistance values of the material used in the
construction of South West (Side Wall).
Figure XI.- Thermal network diagram for South West wall construction
Table VII. Thermal Resistances for South West (Side Wall) material
3.1.2.3 North East (Side Wall). - Table VIII represents the thermal resistance values of the material used in the
construction of North East (Side Wall).
Figure XII.- Thermal network diagram for North East wall construction
Outside Air Resistance
0.044
[25]
Plaster
0.035
[25]
Hollow block
0.172
[29]
Plaster
0.035
[25]
Hardwood
0.141
[25]
Inside Air Resistance
0.120
[25]
Resistance
Description
Value (m2 k/W)
References
Outside Air Resistance
0.044
[25]
Plaster
0.035
[25]
Hollow block
0.172
[29]
Air gap
0.052
[27]
Hollow block
0.172
[29]
Plaster
0.035
[25]
Hardwood
0.141
[25]
Inside Air Resistance
0.120
[25]
Resistance
Description
Value (m2 k/W)
References
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Table VIII. Thermal Resistances for North East (Side Wall) material
3.1.2.4 North West (Stage Wall). - Table IX represents the thermal resistance values of the material used in the
construction of North West (Stage Wall).
Figure XIII.- Thermal network diagram for North West wall construction
Table IX. Thermal Resistances for North West (Stage Wall) Material
Solar time of 1400 hours with Wall Type - E is selected. value is determined from eq (3). Table X depicts
the CLTD value of different walls along with their surface areas.
Table X. Wall Areas & CLTD values
3.1.3 Floor. - The cooling load from the floor can be calculated by eq (4) [28]. U-value is calculated for floor surface
by eq (2). is the temperature below the floor of University Auditorium, taken as 30 °C. The floor area is 615.24 m2.
Table XI represents the thermal resistance values of the material used in the construction of the Floor. Figure XIV
Outside Air Resistance
0.044
[25]
Plaster
0.035
[25]
Hollow block
0.172
[29]
Air gap
0.052
[27]
Hollow block
0.172
[29]
Plaster
0.035
[25]
Hardwood
0.141
[25]
Inside Air Resistance
0.120
[25]
Resistance
Description
Value (m2 k/W)
References
Outside Air Resistance
0.044
[25]
Plaster
0.035
[25]
Hollow block
0.172
[29]
Plaster
0.035
[25]
Inside Air Resistance
0.120
[25]
Wall orientations
CLTD (°C) [28]
Wall Area (m2)
South East
20
67.23
South West
10
168.84
North East
14
168.84
North West
7
103.79
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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represents the thermal network diagram for the construction material of the floor.
(4)
Figure XIV.- Thermal network diagram for Floor Construction
Table XI. Thermal Resistances for Floor Material
3.1.4 Occupants. - The sum of latent and sensible heat is the total heat generated by the people present within the
space. Sensible Heat Gain (SHG) is the heat gained via conduction, convection, or radiation and can be calculated by
eq (5) [28]. If the water molecules are incorporated, then the heat gain in this condition is Latent Heat Gain (LHG) and
can be calculated by eq (6) [28].
(5)
(6)
The values of SHG & LHG are taken as per the desired condition whereas the CLF is selected for 8 hours in space
(0800 hours to 1600 hours) and entry time is taken as 6 hours (from 0800 hours to peak time of 1400 hours). Table XII
depicts the values of SHG & LHG by seated and unseated persons present within the Auditorium.
No. of people who are seated = 550
No. of people who are unseated = 25
Table XII. Sensible & Latent heat gain by a person
3.1.5 Lights. - The lighting load within the space can be calculated by eq (7) [28].
(7)
Resistance
Description
Value (m2 k/W)
References
Inside Air Resistance
0.11
[25]
Plaster
0.035
[25]
Concrete Slab
0.335
[25]
Plaster
0.035
[25]
Outside Air Resistance
0.11
[25]
Load
Seated (W / person)
Unseated
(W / person)
References
Sensible load
60
75
[28]
Latent load
40
75
[28]
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Light is assumed to be operating for 10 hours and the duration of light between turning on and peak time is taken as 6
hours (from 0800 hours to peak time of 1400 hours). Table XIII represents the wattage calculation of lights installed
in the Auditorium while Table XIV depicts the parameters of lighting load calculation.
Table XIII. Wattage of lights
Table XIV. Lighting load parameters
3.1.6 Infiltration. - Infiltration is the uncontrolled introduction of outside air into a building and is represented by .
The ACH (Air Changes per Hour) method is used to determine the ventilation requirement of the Auditorium. ACH
value is taken for medium-construction buildings at an outside temperature condition of 38[30]. V is the volume of
the University Auditorium calculated from the drawing.
(8)
is the sensible heat gain while is the latent heat gain due to infiltration. These values can be calculated by eq
(9) and (10) [28].
(9)
(10)
Table XV. Air Change Method Parameters
3.1.7 Ventilation. - Ventilation air comprises a combination of fresh and recirculated air [28] and plays a crucial role
in ensuring IAQ reaches acceptable levels, meaning the air contains no harmful concentrations of known contaminants
[31]. This system functions by delivering fresh air to indoor areas while concurrently eliminating stagnant air [32].
Fresh outdoor air enters the building through ventilation ducts, where it undergoes filtration and conditioning via a
cooling coil to meet comfort and health requirements. The conditioned air is then dispersed throughout the building
via ductwork and strategically placed vents. Stale air, containing pollutants and excess moisture, is extracted from the
building through exhaust vents, usually directed outside. Some of the air may be recirculated during the exhaust phase
to mix with fresh air before being reintroduced into the space, completing the cycle. This entire process is depicted in
Figure XV.
is the supply air rate for ventilation purposes which can be calculated by eq (11) [28],
is the recirculation air rate
can be calculated by eq (12) [28] and is the minimum outdoor air rate.
(11)
(12)
Type of light
Quantity
Wattage per light
Total Wattage
Energy saver
25
25
625
Recessed light
16
60
960
Total
1590
Parameters
Symbol
Value
Reference
Utilization factor
1
[28]
Ballast factor
1.2
[28]
Cooling Load factor
CLF
0.78
[28]
Wattage
W
1590
--
The volume of the Auditorium ()
ACH [30]
2876.5
0.52
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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is the outdoor air rate and E is the efficiency of contaminant removal by cleaning device. is the sensible heat
gain while is the latent heat gain due to ventilation. These values can be calculated by eq (13) and (14) [28].
(13)
(14)
Table XVI. Ventilation Load Parameters
Figure XV.- Airflow schematic diagram
3.2 Cooling Load estimation using HAP (Transfer Function method)
The software used for the verification of cooling load calculation is Carrier’s HAP version 4.9 as it has a user-friendly
interface, uses the ASHRAE transfer function method for load calculations, and is the fastest way to get a solution and
results rather than doing a manual calculation technique.
The weather properties such as design temperatures and humidity conditions are by default fed in HAP. The only
selection will be of the desired location [34].
The distribution of the desired building into multiple units or sections for thermal comfort is known as Space in which
further parameters like external cooling loads, internal cooling loads, infiltration, and ventilation loads are inserted
[16].
As far as the schedule is concerned, it’s the time and concern day for any activity that is to be performed in any section.
There are some schedules made for lighting purposes, occupants, and other required conditions. Light is assumed to
be operating for 10 hours and the duration of light between turning on and peak time is taken as 6 hours (from 0800
hours to peak time of 1400 hours).
For walls, its construction and structural details for the University Auditorium have been inserted and the R-value for
each material to be used gives out the overall U-value of the wall. Similarly, for the roof, its construction and structural
details have also been inserted, and the R-value for each material to be used gives out the overall U-value of the roof
[35].
The transfer function method (TFM) is especially suitable for computer applications. It provides a straightforward
calculation of the heat gain through a wall or roof, given by the eq (16) [36]:
(15)
3.3 Air Balancing
3.3.1 Fresh / Exhaust air flowrate (Ventilation). - Flow rates of fresh / exhaust air can be determined by eq (15)
[33]. Table XVII represents ventilation rates and related parameters.
Parameters
Values
Reference
2.7
[33]
2.5
[33]
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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(16)
Table XVII. Ventilation rates in the breathing zone
3.3.1.1 Condition of air at different stages. - Concerning the Figure XV, the condition of air at different stages can
be categorized as:
Outside Air - fresh air coming from outside.
Mixed Air - this state is achieved when outside air is mixed with return air from the space.
Supply Air - this state is achieved after mixed air passes through the cooling coil.
Zone Air - this is the state of air inside the conditioned zone.
The supply air flow rate has been taken from HAP software. Eq (17) & (18) are used to find properties at the mixing
state. Table XVIII represents the condition of air at different stages.
(17)
(18)
Table XVIII. Different Parameters for different stages of Air
3.3.2 Supply & Return Air Flowrate (Heat Removal Method). - The supply & return air flow rate is calculated by
eq (19) [28] and eq (20) respectively. Here and are the temperatures at mixed and supply air condition
respectively.
(19)
(20)
3.4 Air Distribution
3.4.1 Diffuser / Return Grill Selection & Duct Routing. - Diffusers are selected based on flow requirements and the
noise criteria permissible for the room [37]. To select a diffuser and return air grill, the flow rate of discharge and
intake air respectively is required. The diffusers and return air grills were selected as per the catalog for price diffusers
[38] and price louvered grilles [39] respectively. Return and Supply air diffusers are divided in such a way that they
provide air to an equal number of outlets. Figure XVI depicts the placement of diffusers and routing of duct from AHU
to Space.
Parameters
Symbol
Values
Units
Reference
People's outdoor air rate
5
cfm / person
[33]
Area outdoor air rate
0.06
cfm / ft2
[33]
No of occupants
N
575
No.
--
Zone area (Area of Auditorium)
--
6619
m2
--
Parameters
Symbol
Values
Units
Total cooling load
202
kW
Sensible load
119
kW
Latent load
82
kW
Fresh air rate
1.54
m3/s
Supply air rate
7.59
m3/s
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Figure XVI.- Duct Routing
3.4.2 Duct Losses. - Duct pressure drop is calculated by using the Equal Friction method. The value of pressure drop
per unit length must be determined from the friction chart [36] and kept constant among all segments. Major loss &
minor loss in pipe segment is calculated by eq (21) & (22) respectively.
(21)
(22)
Table XIX shows the loss coefficient values of some common joints [36].
Table XIX. Loss coefficients
Joint
Loss Coefficient
0.82
0.04
0.73
1.27
0.37
0.52
0.05
16 Pascals
17 Pascals
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Table XX represents the flow, dimensions, and head losses for each segment in the index run of the Supply air duct.
Table XX. Pressure losses in different Supply air duct segments of the Index run
Table XXI represents the flow, dimensions, and head losses for each segment in the index run of the Return air duct.
Table XXI. Pressure losses in different Return air duct segments of the Index run
3.4.4 Duct Sizing. - A rectangular duct for our design (except at the discharge where the cross-section will be circular)
is selected, equivalent diameter is calculated by eq (23) [28].
(23)
3.5 Equipment Selection. - We are proposing two different HVAC systems for the University Auditorium which are
based on water cooled chiller and an air-cooled chiller.
3.5.1 Proposed System # 1. - This system primarily relies on a water-cooled chiller. Its main components consist of a
compressor, condenser, expansion valve, and evaporator. Chilled water from the evaporator is directed to the AHU via
two chilled water pumps (one operational and the other on standby). The AHU comprises various elements, including
filters, a supply air duct, a fresh air duct, a return air duct, an air mixing chamber, and a cooling coil. Chilled water
Major
Minor (KL)
No. of Fittings
Segment
(Pa/m)
Flow
rate
(L/s)
Velocity (m/s)
Length (m)
Entrance
Wye - through
Wye - branch
Tee - through
Tee - branch
K90
K45
A
0.15
8307
5.0
4.93
1
--
--
--
--
--
--
B
0.15
4153
4.2
13.47
--
1
--
--
--
1
--
B1
0.15
4153
4.2
10.01
--
--
--
--
--
1
--
D
0.15
4153
4.2
15.24
--
--
--
--
--
2
--
G
0.15
2076
3.6
6.83
--
1
--
--
--
--
--
I
0.15
1038
3.0
4.85
--
1
1
--
--
1
--
I1
0.15
519.2
2.5
0.30
--
--
--
--
--
--
--
Major
Minor (KL)
No. of Fittings
Segment
(Pa/m)
Flow
rate
(L/s)
Velocity (m/s)
Length (m)
Entrance
Wye - through
Wye - branch
Tee - through
Tee - branch
K90
K45
A
0.08
6570
3.8
0.58
1
--
--
--
--
--
--
C
0.08
3285
3.3
5.08
--
--
--
--
--
1
--
G
0.08
2190
2.8
7.85
--
--
--
--
1
1
--
G1
0.08
2190
2.8
11.64
--
--
--
--
--
1
--
K
0.08
1095
2.4
1.04
--
--
--
--
--
1
--
K1
0.08
1095
2.4
1.65
--
--
--
--
--
1
--
K2
0.08
1095
2.4
10.98
--
--
--
--
--
1
--
K3
0.08
1095
2.4
0.66
--
--
--
--
--
1
--
K4
0.08
1095
2.4
1.65
--
--
--
--
--
1
--
K5
0.08
1095
2.4
0.66
--
--
--
--
--
1
--
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flows through the coil while air passes over it, becoming cooled in the process. The cooled air is then distributed into
the space. Another essential part is the condenser water circuit. Hot water from the condenser is conveyed to the cooling
tower through two condenser water pumps (one operational and the other on standby). After heat rejection, the water
returns to the condenser, completing the cycle. The term "water-cooled" refers to the condenser water being cooled by
the cooling tower. The entire system is illustrated in Figure XVII for reference.
Figure XVII.- Proposed System # 1
3.5.2 Proposed System # 2. - This system operates on an air-cooled chiller. Its main components are identical to those
of a water-cooled chiller, except for the absence of a cooling tower. Instead, an air-cooled condenser is used,
eliminating the need for a cooling tower as the condenser water is cooled naturally by air. This system configuration
is depicted in Figure XVIII.
Figure XVIII.- Proposed System # 2
3.5.3 Chiller & AHU selection. - Typically, the chiller supplies chilled water at around 7 °C which makes the return
water temperature 12 °C. The chilled water flow rate will be taken from water cooled chiller’s datasheet. The supply,
return, and fresh air flow rates have been taken from the Air Balancing section. The fan selection performed in the
AHU is based on the external and internal pressure drops. Internal pressure drop is due to the losses inside AHU while
external pressure drop is the sum of pressure drop in the damper, ducting, and diffusers.
3.5.4 Cooling Tower & Pump selection. - The condenser water flow rate will be taken from the chiller’s datasheet.
Cooling tower inlet and outlet water temperatures depend upon the selection of chiller because the condenser’s inlet
and outlet water conditions vary with different chillers. For wet bulb temperature, outside air data will be used. We
will select two pumps each for the condenser and chilled water circuits. One pump is working while the other is for
backup. The cooling tower’s height, condenser & evaporator losses (from the chiller datasheet), pipes, fittings, and
valve losses will be considered for the head. The flow rate will be taken from the selected water-cooled chiller
datasheet.
3.6 Piping System
3.6.1 Water Cooled System. - For the application of HVAC, steel pipes are commonly used. We are selecting steel
pipe of Schedule 40 for condenser and chilled water piping. For pipe sizing, we are using Carrier’s friction charts [36]
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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for Schedule 40 Pipe. Table XXII depicts the Pipe design parameters for the cool and chilled water circuit.
Table XXII. Pipe design parameters for Cool and chilled water system
The total piping length in the condenser and chilled water system is 161 ft & 79 ft respectively. Head losses due to
valves and fittings for condenser and chilled water lines are taken [40] for 4” & 3.5” pipe diameters respectively. Table
XXIII shows the equivalent lengths of the head, no. of valves, and fittings in the condenser and chilled water line.
Table XXIII. Pipe fittings, valves & equivalent length in condenser & chilled water line
Head loss due to piping is given by eq (24) [40].
(24)
The pressure loss in the condenser and evaporator is calculated by eq (25) [40].
(25)
3.6.2 Air Cooled System. - Steel pipe of Schedule 40 for Chilled Water Piping is selected. For pipe sizing, we select
Schedule 40 Pipe [40]. Table XXIV depicts the Pipe design parameters for the chilled water circuit.
Table XXIV. Pipe design parameters for chilled water system
The total piping length in a chilled water system is 112 ft. Head losses due to valves and fittings for chilled water lines
are taken [40] for a 3.5” pipe diameter. Table XXV depicts the equivalent lengths of the head, no. of valves, and fittings
in the chilled water line.
Parameters
Condenser water
Chilled water
Unit
Ref
Velocity
5
6
Ft/sec
[40]
Flowrate
205
156
US gpm
[40]
Pipe size
4
3.5
inch
[40]
Friction loss
3.6
2.5
Ft of water per 100 ft
[40]
Designing Criteria
Condenser water line
Chilled water line
L/D
Quantity
Leq (ft)
L/D
Qty
Leq (ft)
Fittings
Elbow 90°
10
20
66.7
9
16
48
Tee (flow thru branch)
21
4
28
18
4
24
Valves
Gate
4.5
4
6
4
3
4
Globe
120
2
80
100
1
33.3
60° Strainer
60
1
20
48
1
16
Swing Check
40
1
13.3
35
1
11.7
Total
214
137
Total equivalent length
375
216
Parameters
Chilled water
Unit
Reference
Velocity
6
Ft/sec
[40]
Flowrate
168
US gpm
[40]
Pipe size
3.5
inch
[40]
Friction loss
3.6
Ft of water per 100 ft
[40]
Designing Criteria
Chilled water line
L/D
Quantity
Leq (ft)
Fittings
Elbow 90°
9
20
60
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Table XXV. Pipe fittings, valves & equivalent length in a chilled water line
3.6.3 Duct & Piping Layout. - Figure XIX & XX shows the ducting and piping layouts for the two proposed HVAC
systems for the Auditorium while Figure XXI depicts a ducting layout of the plant room for both the proposed HVAC
systems.
Figure XIX.- Duct & Pipe Layout for Proposed System # 1
Tee (flow thru branch)
18
4
24
Valves
Gate
4
3
4
Globe
100
1
33.3
60° Strainer
48
1
16
Swing Check
35
1
11.7
Total
149
Total equivalent length
261
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Figure XX.- Duct & Pipe Layout for Proposed System # 2
Figure XXI.- Duct Layout of Plant Room for Proposed Systems # 1 & 2
4. Results. - In this study, we calculated the cooling load of the University Auditorium first by manual calculation
using the CLTD method then the results obtained were validated by using CARRIER’s HAP (version 4.9) software.
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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4.1 Cooling Load by CLTD Results. - The results of the cooling load obtained after calculation by using CLTD
methods are listed in Table XXVI. All the cooling loads are taken in kW.
Table XXVI. Cooling Load (CLTD method)
The cooling load obtained from the CLTD-based calculation was 202 kW which is equivalent to around 57 TR.
Figure XXII shows the condition of air at different stages on a psychrometric chart calculated by the CLTD method.
Figure XXII.- Psychrometric Chart (CLTD Results)
4.2 Cooling Load by HAP Results
The HAP result depicts the cooling load of the University Auditorium as 192.8 kW which is equivalent to around
55 TR.
The calculated ventilation (fresh air) rate for desired air conditioning in the air balancing section is 3273 cfm which
is close to the value determined from the HAP result which is 1622 L/s (3436 cfm).
The calculated supply air flow rate from the Heat Removal method is 16071 cfm which is close to the value given
by the HAP result as 7589 L/s (16080 cfm).
External Cooling Load
Roof
23.96
Walls
South East (Back wall)
3.7
South West (Sidewall)
4.4
North East (Sidewall)
5.28
North West (Stage wall)
4.37
Floor
7.87
Internal Cooling Load
Occupant
51.78
Lights
1.48
Infiltration Load
20.98
Ventilation Load
78.38
Total Cooling Load of the Auditorium
202
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Figure XXIII & XXIV shows the results obtained by HAP.
Figure XXIII.- HAP Results
Air System Sizing Summary for University
Auditorium
Project Name: University
Auditorium
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Figure XXIV.- Sensible & Latent heat gains
Figure XXV depicts the HAP result which shows a psychrometric diagram of the Air conditioning process for the
University Auditorium.
Figure XXV.- Psychrometric chart obtained from HAP software
Air System Design Load Summary for University
Auditorium
Project Name: University
Auditorium
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Figure XXVI shows the comparison of the University Auditorium cooling loads between CLTD and HAP. Figure
XXVII depicts the comparison of Total Heat Gain by CLTD & HAP software.
Figure XXVI.- Cooling load (CLTD vs HAP)
Figure XXVII.- Total Heat Gain (CLTD vs HAP)
4.3 Air Balancing. - Figure XXVIII shows the Fresh air, return air & Supply air flow rates by using the CLTD method.
Figure XXVIII.- Air Balancing Flowrates
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4.4 Diffuser & Return Air Grill Selection. - A total sixteen number of diffusers in the University Auditorium have
been calculated. We have selected Round Diffusers [38] because they meet the low NC requirement for the Auditorium.
The diffuser that meets our requirement is Price’s 20” Round Cone Diffuser (RCD) providing supply air of 1100 cfm
at a minimal NC value of 15. Air velocity is 500 fpm whereas the throw at 150 mm is 14 ft.
Similarly, we will install six return air grills in the Auditorium space and select Price’s 40” x 16” Rectangular Louvered
Grille, whose intake flow rate is 2320 cfm [39].
4.5 Duct Losses. - Table XXVII represents the results obtained for the major and minor losses through different Supply
air duct segments of the Index run. Table XXVIII represents the results obtained for the major and minor losses through
different Return air duct segments of the Index run.
Table XXVII. Supply air duct losses
Table XXVIII. Return air duct losses
4.6 Duct Sizing. - Supply & return air duct sizing results for an aspect ratio of 1.5 are shown in Table XXIX & Table
XXX respectively.
Segment
Major Loss (Pa)
Minor Loss (Pa)
Head Loss (Pa)
Total Pressure Drop (Pa)
A
0.72
12.7
--
13.43
B
1.98
14.43
--
16.41
B1
1.47
13.88
120.29
135.65
D
2.24
27.76
--
30.00
G
1.00
0.39
--
1.39
I
0.71
7.36
--
8.07
I1
0.04
2.01
--
2.06
Pressure drop across supply diffuser
16
Total
223.01
Segment
Major Loss (Pa)
Minor Loss (Pa)
Total Pressure Drop (Pa)
A
0.05
7.34
7.39
C
0.41
8.31
8.73
G
0.64
9.37
10.01
G1
0.95
5.95
6.90
K
0.09
4.35
4.43
K1
0.13
4.35
4.48
K2
0.90
4.35
5.24
K3
0.05
4.35
4.40
K4
0.13
4.35
4.48
K5
0.05
4.35
4.40
Pressure drop across the return grill
17
Total
77.46
Segment
Dia. (cm)
Size
Duct Shape
a (cm)
b (cm)
A
144
164
109
Rectangular
B
111
126
84
Rectangular
B1
111
126
84
Rectangular
C
111
126
84
Rectangular
C1
111
126
84
Rectangular
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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Table XXIX. Supply air duct sizing
D
111
126
84
Rectangular
E
66
75
50
Circular
E1
51
59
39
Circular
E2
51
59
39
Circular
F
66
75
50
Circular
F1
51
59
39
Circular
F2
51
59
39
Circular
G
86
98
65
Rectangular
H
66
75
50
Circular
H1
51
59
39
Circular
H2
51
59
39
Circular
I
66
75
50
Circular
I1
51
59
39
Circular
I2
51
59
39
Circular
J
111
126
84
Rectangular
J1
51
59
39
Circular
J2
106
120
80
Rectangular
K
93
105
70
Rectangular
L
66
75
50
Circular
M
51
59
39
Circular
N
51
59
39
Circular
O
51
59
39
Circular
P
51
59
39
Circular
Q
77
89
59
Circular
R
51
59
39
Circular
S
51
59
39
Circular
T
51
59
39
Circular
Fresh
82
93
62
Rectangular
Segment
Dia. (cm)
Size
a (cm)
b (cm)
A
147
166
110
B
112
127
84
C
112
127
84
D
76
86
57
D1
76
86
57
D2
76
86
57
E
76
86
57
E1
76
86
57
E2
76
86
57
F
100
112
75
F1
100
112
75
G
100
112
75
G1
100
112
75
H
76
86
57
I
76
86
57
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Table XXX. Return air duct sizing
4.7 Chiller Selection
4.7.1 Proposed System # 1. - As per the design requirement and the consultant’s provided design condition datasheet,
the chiller we have selected is YORK’s 65-ton scroll compressor-type water-cooled chiller.
Table XXXI. Specification of Water-cooled chiller
This chiller consists of four scroll compressors which provide a part load condition. All four compressors are operating
at maximum cooling load conditions but as the cooling requirement of the zone decreases, the system automatically
turns off one of the compressors. Therefore, the number of compressors in operation depends on the zone cooling
requirement resulting in saving power. Moreover, even at the maximum capacity, the system operating power is still
lower than the existing installed system whose compressor’s operating power is 55 kW. GWP of R-410 (A) is 2088.
4.7.2 Proposed System # 2. - Another chiller that is nearest to our design capacity and as per the consultant’s provided
design condition datasheet is YORK’s 71 tons Air-cooled Scroll Chiller.
Table XXXII. Specification of Air-cooled chiller
This chiller consists of six scroll compressors, R-410 (A) as a refrigerant, and works on the same part load condition
as water cooled chiller does. As compared to water cooled chiller, this chiller has a lower value of energy efficiency
and its power input is more than double that of the water-cooled chiller. The chiller has an approx. size of 3400 mm x
J
76
86
57
J1
76
86
57
J2
76
86
57
J3
76
86
57
J4
76
86
57
J5
76
86
57
K
76
86
57
K1
76
86
57
K2
76
86
57
K3
76
86
57
K4
76
86
57
K5
76
86
57
YORK’s 65 tons scroll compressor type water-cooled chiller specification
Model
YCWL0261HE
Refrigerant
R-410 (A)
EER
14.90 (Btu/W∙h)
Chilled water supply/return temp.
7 °C / 12 °C (44 °F / 54 °F)
Condenser water supply/return temp.
32 °C /38 °C (90 °F /100 °F)
Condenser / chilled water flow rates
205 / 156 US gpm
Condenser/evaporator pressure drop
8.01 / 5.74 ft. of water
Maximum operating power
52 kW
YORK’s 71 tons Air-cooled Scroll Chiller specification
Model
YLAA0286SE
Refrigerant
R-410 (A)
EER
7.47 (Btu/W∙h)
Chilled water supply/return temp.
7 °C / 12 °C (44 °F /54 °F)
Chilled water flow rates
168 US gpm
Pressure loss
10.6 ft. of water
Maximum operating power
115 kW
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2250 mm x 2400 mm. The roof of the University Auditorium where the current cooling tower is installed has sufficient
space to accommodate this chiller.
4.8 AHU Selection. - Table XXXIII depicts the important parameters for selection of AHU which we have previously
calculated.
Table XXXIII. AHU selection parameters
By considering the above parameters and as per the consultant’s provided datasheet, we have selected AHU from
YORK, having Model # YMA(T)1730H-2450W. This AHU provides a supply air flow rate of 8.30 m3/s (17600 cfm
approx.) having a size of 3000 mm x 2450 mm x 1830 mm (smaller than the AHU of an existing system whose
dimensions are 3940 mm x 3090 mm x 1740 mm) and will easily be fitted inside the plant room. The fan must overcome
a total pressure drop of 750 Pa. (internal Pressure Drop = 277 Pa, external Pressure Drop = 473 Pa). For this purpose,
a fan of 15 kW of nominal power is selected. This AHU model is applicable for both proposed System # 1 and System
# 2.
4.9 Cooling Tower Selection. - Table XXXIV depicts the important parameters for the selection of a cooling tower
which we have previously calculated.
Table XXXIV. Cooling tower selection parameters
By considering the above parameters and as per the consultant’s design condition datasheet, we have selected Liang
Chi’s bottle type counter flow induced draft cooling tower of Model # LBC-60-S. This model can provide a flow rate
at 205 US gpm and meets the required condenser temperatures. It has a direct motor drive axial-flow fan that runs at
750 rpm with a motor input power of 2 hp. Furthermore, the inlet and outlet connection have a 3” dia. so a reducer
must be required for the installation of the condenser water pipe. The roof of the University Auditorium is a suitable
place for its installation.
4.10 Pump Selection. - Table XXXV depicts the important parameters for the selection of Chilled/condenser water
pumps which we have previously calculated.
Table XXXV. Pump selection parameters
AHU Selection
Supply air flow rate
16071 cfm
Return air flow rate
12800 cfm
Fresh air flow rate
3272 cfm
Off-coil temperature (dbt / wbt)
26 °C / 20 °C
On coil temperature (dbt / wbt)
13 °C / 11 °C
External press. drop
223 x 1.25 (Factor of safety) = 277 Pa
Chilled water flow rate
156 / 168 US gpm
Chilled water supply/return temperature
7 °C / 12 °C
Cooling Tower Selection
System capacity
65 tons
Condenser water flow rate
205 US gpm
Cooling tower inlet/outlet temperatures
32 °C / 38 °C (90 °F /100 °F)
Design wet bulb temperature
23.2 °C (73 °F)
Parameters
Unit
Water cooled system
Air-cooled system
Chilled water pump
Condenser water pump
Chilled water pump
Flowrate
US gpm
156
205
168
Pressure head
Ft
32.43
36.37
51.26
Pipe dia.
inch
3.5
4
3.5
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By considering the above parameters and as per the consultant’s design condition datasheet, we have selected KSB’s
low-pressure centrifugal pumps. Table XXXVI depicts the model numbers and quantity of selected chilled & cool
water pumps for both proposed systems.
Table XXXVI. Pumps Model Number
These models meet our required head and flow rates. Motor input power and speed of chilled water and condenser
water pumps are 2.95 hp each & 1400 rpm for water cooled system. The motor input power and speed of the chilled
water pump are 5.36 hp & 1400 rpm for the air-cooled system.
4.11 Piping System. - Table XXXVII depicts the head loss from the water-cooled and air-cooled chilled & cool water
piping system. All losses are measured in ft.
Table XXXVII. Piping head loss
4.12 Piping Layout. - Figure XXIX & XXX depicts the piping layout of the plant room for both the proposed HVAC
systems.
Figure XXIX.- Pipe Layout of Plant Room for Proposed System # 1
System
Pump
Quantity
Model No.
Water cooled system
Chilled Water Pump
2
ETN 065-050-200 GBSAA11GD200224B
Condenser Water Pump
2
ETN 080-065-200 GBSAA11GD200224B
Air-cooled system
Chilled Water Pump
2
ETN 080-065-200 GBSAA11GD200404B
Head Loss
Water cooled system
Air-cooled system
Condenser water circuit
Chilled water circuit
Chilled water circuit
Pipe Friction
15.72
7.19
18.01
Condenser / Evaporator
8.03 / --
-- / 5.75
-- / 10.62
Cooling Tower / AHU
6.56 / --
-- / 14.09
-- / 14.09
Total Head Loss (with 20%
Factor of Safety)
36.4
32.43
51.26
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Figure XXX.- Pipe Layout of Plant Room for Proposed System # 2
5. Conclusion & Discussion. - This study outlines the design of HVAC systems for the University Auditorium,
proposing both water-cooled and Air-Cooled Vapor Compression systems. Cooling loads were calculated using the
CLTD method and HAP software, resulting in 57 TR and 55 TR respectively, with a small discrepancy of only 3.5%.
The fresh air flow rate was manually calculated at 3272 cfm, closely matching HAP's 3436 cfm, while for the Supply
air flow rate, the manual calculation (Heat Removal method) yielded 16071 cfm compared to HAP's 16080 cfm.
Equipment selections include sixteen Price’s 20-inch Round Cone Diffusers, six Price’s 40” x 16” Rectangular
Louvered Grille, YORK’s 65 tons scroll compressor type water-cooled chiller, YORK’s 71 tons Air-cooled Scroll
Chiller, YORK’s AHU, and Liang Chi’s bottle type counter flow induced draft cooling tower, along with two KSB’s
centrifugal pumps. The study highlights the importance of designing cost-effective and energy-efficient HVAC
systems, addressing the increasing demand driven by global warming and humid climates, and offering guidance for
engineers and researchers to propose new systems for spaces, including replacing obsolete ones, based on selection
criteria and research findings.
A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
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A. Samad Khan, M. Ehtesham ul Haque, A. Ahmed Khan, S. Izhar ul haque, S. Obaidullah, M. Umer Khan
Memoria Investigaciones en Ingeniería, núm. 26 (2024). pp. 2-37
https://doi.org/10.36561/ING.26.2
ISSN 2301-1092 • ISSN (en línea) 2301-1106 – Universidad de Montevideo, Uruguay 37
Nota contribución de los autores:
1. Concepción y diseño del estudio
2. Adquisición de datos
3. Análisis de datos
4. Discusión de los resultados
5. Redacción del manuscrito
6. Aprobación de la versión final del manuscrito
ASK ha contribuido en: 1, 2, 3, 4, 5 y 6.
MEUH ha contribuido en: 1, 2, 3, 4, 5 y 6.
AAK ha contribuido en: 1, 2, 3, 4, 5 y 6.
SIUH ha contribuido en: 1, 2, 3, 4, 5 y 6.
SO ha contribuido en: 1, 2, 3, 4, 5 y 6.
MUK ha contribuido en: 1, 2, 3, 4, 5 y 6.
Nota de aceptación: Este artículo fue aprobado por los editores de la revista Dr. Rafael Sotelo y Mag. Ing. Fernando
A. Hernández Gobertti.
