Sensible and latent heat from infiltration air Q5...21 2.2.10 Heat loss from other sources Q6...22 2.3 Establish and calculate the air conditioning system diagram...23... α k = 0.25; τ
PROJECT OVERVIEW
Project Introduction
Project name: “NATIONAL UNIVERSITY LIBRARY”
Location: Linh Trung Ward, Thu Duc City, Ho Chi Minh City.
Project Owner: CONSTRUCTION PROJECT MANAGEMENT BOARD –
VIETNAM NATIONAL UNIVERSITY HO CHI MINH CITY.
Design Consultant Unit:LAP VIET CONSTRUCTION INVESTMENT CONSULTING
Project Features
The Central Library of Vietnam National University Ho Chi Minh City covers about 8,600 m² and is organized across an upper ground floor, a ground floor, three upper levels, and a mezzanine It houses key functional zones for study and learning, including large reading halls, quiet study rooms, group-study areas, multimedia spaces, computer labs, administrative offices, and storage areas On the ground floor, visitors are welcomed by a lobby and a multifunctional exhibition space, along with a 24/7 study area, a book return and classification zone, a circulation desk, and a café The upper floors feature an automatic borrowing–return system, book stacks, open reading spaces, meeting rooms, studios, VR rooms, digital resource areas, and a 100-seat auditorium The top levels contain archives, technical rooms, and document storage.
The library at Vietnam National University Ho Chi Minh City features modern facilities and a state-of-the-art digital infrastructure that integrates leading technologies for library management, digital cataloging, user authentication, and fast information search It offers flexible learning spaces and high-tech services designed to support a wide range of academic activities—from independent study and group collaboration to research, training workshops, exhibitions, and community engagement—establishing a modern, innovative learning hub within the university.
Scope of the study
Within the renovation and upgrade project for the air-conditioning and ventilation systems at the VNU-HCMC Central Library, this study concentrates on the main tasks required to modernize and optimize performance, including evaluating the current infrastructure, identifying improvement opportunities, defining system requirements for enhanced comfort and indoor air quality, and planning the implementation to ensure energy efficiency, reliability, and minimal disruption to library operations.
An HVAC system assessment was conducted in strict alignment with the official design specifications, detailing the renovation scope to replace the AC electrical panel, install a centralized VRV/VRF system, upgrade fresh-air and return-air handling, establish condensate drainage, and implement centralized control systems The evaluation identifies current performance gaps, assesses compatibility with VRV/VRF technology, and outlines integration with new air distribution and energy-management strategies to improve comfort, efficiency, and reliability The renovation plan calls for electrical panel replacement to satisfy load requirements, installation of a modular VRV/VRF plant, dedicated fresh-air and return-air systems, robust condensate drainage, and a centralized control interface for monitoring and operation This results in a clear implementation roadmap, including commissioning and testing procedures aligned with design standards.
Conducting cooling load calculations according to the project’s design criteria (indoor temperature 24 ± 1°C, humidity 55 ± 5%, occupant density, lighting load, glazing heat gain, ) to verify the required cooling capacity.
Assessing and verifying the ventilation system, including toilet exhaust, technical room and storage ventilation, and outdoor air supply for functional spaces as specified in the design document.
Evaluating and verifying the pressurization and smoke extraction system, in accordance with fire protection requirements and the official approvals from authorities related to the project.
Developing schematic diagrams of the HVAC system and verifying major components, such as VRV/VRF outdoor–indoor units, PAUs, fresh air and return air systems, refrigerant piping,
AC electrical panels, and central control units.
Creating a 3D Revit MEP model of the air-conditioning and ventilation systems, including equipment layout, ducting design, condensate piping, and generating a visual animation of the system.
COOLING LOAD CALCULATION
Fundamentals for the design and performance evaluation of HVAC and ventilation systems
2.1.1 Introduction of HVAC system of the project
VNU-HCMC Central Library, a public building, houses diverse functional spaces such as large reading rooms, group-study rooms, 24/7 study areas, functional rooms, exhibition spaces, administrative offices, and technical rooms Each area presents different cooling loads and operating schedules, making small-capacity air conditioners unable to meet the building’s operational requirements To address this, a centralized VRV/VRF air-conditioning system was selected to deliver cooling efficiency, operational flexibility, and energy savings across the facility.
The newly installed VRV/VRF central air-conditioning system places rooftop outdoor units to optimize technical space and preserve architectural aesthetics, while the indoor lineup combines cassette-type, ducted concealed, and wall-mounted units chosen by room function to maximize comfort and efficiency This configuration provides uniform climate control and consistent thermal comfort throughout the building.
Additionally, the HVAC system integrates fresh air and return air systems, condensate drainage, new AC electrical panels, and a centralized control system, along with ventilation fans for toilets, technical rooms, and storage rooms, all designed in accordance with TCVN 5687:2024 and ASHRAE 62.1 standards.
This project is builed in Ha Noi, so we have the parameters of climate for this project:
Look up the t – d chart we have:
Indoor conditions vary by area, since each space has a different function, which means both indoor temperature and humidity levels differ In most cases, two key heat parameters—temperature and humidity—are used to characterize the indoor climate The specific temperature and humidity values are determined by the intended use of the space and are guided by the source cited as [2].
Zone 1: Reading rooms, group-study rooms, functional rooms, staff rooms:
Look up the t – d chart we have:
Zone 2: Lobby, corridors, circulation areas
Look up the t – d chart we have:
Calculating cooling load using the Carrier method
The cooling load calculation using the Carrier method differs from the traditional method in that it determines the cooling capacity Q 0 in summer and the heating capacity
Q s in winter by separately calculating the sensible heat Q ht and latent heat Q at generated and transmitted into the air-conditioned space
Heat emitted from lighting equipment and machinery Q3.
Sensible and latent heat emitted by humans Q4.
Sensible heat and latent heat brought into QN by the fresh wind.
2.2.1 Sensible radiant heat gain through glass Q 11
Sensible radiant heat gain through glass Q11 is determined by the following equation from [1] :
Q′11: The instantaneous radiant heat gain through the glass into the room is determined using equation from [1]:
Fk: Area of glass surface (m2).
RT: Solar radiation heat gain through the glass into the room (W/m2).
c : The coefficient of the influence of altitude relative to sea level.
ds: The coefficient of the influence of the difference between the observed air dew point temperature and the standard dew point temperature at sea level (20℃).
mm : Coefficient accounting for the effect of fog and clouds.
kh : Coefficient considering the influence of the frame.
Determine the coefficient of influence
The coefficient of the influence of altitude relative to sea level c is calculated according to formula:
With the natural ground elevation above sea level being 9m
1000 = 1.000207For the convenience of calculation, we choose c = 1.
The correction factor ε_bd accounts for the difference between the observed air dew point temperature and the standard sea-level dew point of 20°C It is determined by the formula ε_ds = 1 − 0.13(t_s − 20), where t_s is the observed dew point temperature.
10 ts: Dew point temperature of the outdoor air (oC).
With t N = 36.1 ℃ and φ N = 50.1% , consult the t-d graph we have ts = 24.06 ℃ ε ds =1−0.13 t s − 20
-The coefficient accounting for the effect of fog and clouds mm:
For clear skies mm = 1; for the cloudy skies mm = 0.85.
Since HoChiMinh climate is often cloudy, we choose mm = 0.85.
-The coefficient considering the influence of the frame kh
Different frame structures cause varying degrees of partial obstruction of the glass under different radiation rays For a wooden frame, ε kh = 1; for a metal frame, εkh = 1.17.
Since the window frames are made of metal, the frame effect coefficient εₖₕ is taken as 1.17.
- The solar factor ε r This coefficient considers the influence of shading devices on solar radiation As there are no shading devices used in the building, ε r =1.
- The glass coefficient ε : The external glazing used in the project consists of laminated glass, total α k = 0.25; τ k = 0.6; k = 0.15; m = 0.84
VNU-HCMC Central Library sits at about 10.75°N latitude in Ho Chi Minh City The city’s hottest month is April, with average temperatures around 29–30°C, while the outdoor design dry-bulb temperature is typically taken as 35–36°C according to regional climatic standards.
Table 2.1: Coefficient RTmax (W/m2) of each side:
Q 11 :is the amount of heat that directly contributes to the cooling load
Q 11 ' : is the instantaneous radiant heat passing through the glass into the room but does not directly contribute to the cooling load
G′: Mass of walls with external surfaces exposed to solar radiation and of floors in contact with the ground (kg)
G″: Mass of walls with external surfaces not exposed to solar radiation and of floors not in contact with the ground (kg)
The unit mass of glass is 2500 kg/m³ The exterior glass of the building is 6 mm thick and has an area of 619 m² We have:
The unit mass of the reinforced concrete floor is 2400 kg/m³ The concrete floor of the building is 250 mm thick and has an area of 691 m² We have:
We calculate the average area mass density as: g s = G
So we have g s > 150 kg/ m 2 , refer to table 4.6 – [1] we have the following table 2.2
Table 2.2: Instantaneous action coefficient nt of each side
For areas with glass facing Northeast:
For areas with glass facing East:
Table 2.3: Heat gain of radiation through glass of level 2
Room/Zone Direction Glass area
Sound & Lightning control room East
2.1.2 Sensible heat transfer through the roof by radiation and ∆T: Q 21
According to [1], the flat roof of an air-conditioned room can fall into three categories:
If the air-conditioned room is located between floors of an air-conditioned building meaning the space above is also air-conditioned, then Δt = 0 and Q 21 = 0.
If the space above the room being calculated is not air-conditioned, then the heat transfer coefficient k is taken from table 4.15 [1], and Δt = 0.5(tN − tT).
In the case where the ceiling is directly exposed to solar radiation as in the top floor of a multi-story building the heat transfer into the room includes two
2.1.3 Sensible heat gain through walls Q 22
According to [1], the heat transfer through the wall Q22 consists of two components:
Due to the temperature difference between the outdoor and indoor environments, Δt = tN − tT.
Due to solar radiation on the wall; however, this component is considered negligible in the calculation.
The sensible heat gain through walls Q22 is determined by the following expression:
Q2i: Heat transfer through walls, doors (wood, aluminum), windows (glass),…, (W)
Q₂₂t:Heat transfer through the wall, (W)
Q₂₂c: Heat transfer through the door,
(W) Q₂₂k: Heat transfer through the glass, (W) kᵢ: Heat transfer coefficient of the wall, door, and window glass, (W/m²ãK)
Fᵢ: Surface area of the wall, door, and glass, (m²) Δt: Temperature difference between the outdoor environment and the conditioned indoor space, (°C)
When the wall is in direct contact with the outdoor environment:
When the wall is adjacent to an unconditioned space inside the building:
The heat transfer coefficient of the wall is determined by the following equation:
+ 1 1 δv δg 1 α N λ i α T αN + λv + λg + αT δv, v : Thickness and thermal conductivity of the plaster layer δg, g : Thickness and thermal conductivity of the standard brick layer.
N = 20 (W/m2K): External convective heat transfer coefficient (when the wall is exposed to outdoor air).
T = 10 (W/m2K): Internal convective heat transfer coefficient (on the indoor side).
Heat transfer coefficient of the wall when directly exposed to the outdoor environment:
20 0.81 0.93 10 Heat transfer coefficient of the wall when adjacent to an unconditioned space:
When the wall is adjacent to an unconditioned space, the calculation is similar to the previous case, except that α = 10 W/m²Kₙ
Table 2.4: Heat transfer through walls Q22t of level 2
Heat transfer through door Q22c is determined using the following equation:
In which: kc (W/m²ãK): Heat transfer coefficient of the door, determined according to Table 4.12 [1]
Fc (m²): Area of the door Δt: Temperature difference between indoor and outdoor environments
As all entrance doors of the rooms on the 2 nd floor are in contact with an air- conditioned corridor, the corresponding heat gain is considered to be zero.
Heat transfer through glass window Q 22k
Heat transfer through glass window Q22k:
Q 22k = kkFk∆t (W) (2.12) kk (W/m²ãK): Heat transfer coefficient of the glass window
Fc (m²): Area of the glass window k = 3.15 (W/m2K) – tra theo bảng 4.13 [1]
t: Temperature difference between external and in air conditioning space
t = (tN – tT) = (33.4 – 22) = 11.4oC (With indoor temperature of 22oC)
t = (t N – tT) = (33.4 – 26) = 7.4 oC (With in door temperature of 24oC)
Table 2.5: Heat transfer through glass window Q22k of level 2
2.1.4 Sensible heat gain through floor Q 23
Sensible heat gain through floor Q23 is determined using the following equation:
F: floor area (m²) k: heat transfer coefficient (W/m²ãK), determined from Table 4.15 [1] Δt: temperature difference between outside and inside, Δt = tₙ – tₜ (°C) The following cases apply:
For floors in contact with the ground: Δt = tN – tT
For floors above unconditioned spaces: Δt = 0.5 × (tN – tT)
For floors above conditioned rooms: Q₂₃ = 0
Since the entire third floor is a conditioned space, Q23 = 0 is applied in this case The cooling load is smaller than the total calculated heat The sensible heat emitted by lighting, denoted as Q31, is determined using the equation below.
Lighting calculations use nt, the instantaneous utilization factor of the light For areas with gs > 150 kg/m2, consult Table 4.8 in ĐHKK – Nguyen Duc Loi to obtain nt, which equals 0.98 The simultaneity factor nd, primarily used for houses and large air-conditioning buildings, is set to nd = 0.85 for the theater, a large building, based on reference [1] Finally, qs denotes the lighting power density required per square meter of floor space.
Under the growing emphasis on energy savings, most large buildings now rely on LED lighting When the design assumes LED-only operation, a correction factor of 1.25 is applied, as derived from Equation 4.14 in reference [1] Accordingly, the final formula for calculating the sensible heat from lighting, denoted as Q₃₁, is obtained using this corrected relationship, linking LED usage directly to the building’s cooling load This method provides a robust, SEO-friendly way to estimate lighting heat gains and support efficient HVAC design in modern, energy-conscious facilities.
According to Table 9.6.1 – [3]/Table 2 - chapter 18 – [4], we have table 2.6:
Table 2.6: Lighting power density of level 2
So sensible heat gain from lighting Q31 of level 2 is caulated in the table 2.7.
Table 2.7: Sensible heat gain from lighting Q31 of level 2
2.1.6 Sensible heat gain from equipment Q 32
Q32 describes the heat emitted by common electrical appliances and devices used in homes and offices, such as hair dryers, televisions, irons, and computers Even devices without electric motors generate thermal energy, making them heat sources that are comparable to lighting fixtures Recognizing these heat sources helps assess indoor heat load, energy use, and the overall thermal environment of occupied spaces.
Ni: electrical power written on the device (W)
Since the team cannot survey the actual construction site, they rely on architectural drawings to estimate the number of equipment expected on site, using the office as a reference for the equipment count Based on this estimate, they calculate the corresponding heat output that the equipment would emit during construction, enabling an early thermal assessment before on-site inspection.
Table 2.8: Equipment power density of level 2
Sensible heat gain from equipment Q32 of level 2 is caculated in table 2.9:
Table 2.9: Sensible heat gain from equipment Q32 of level
2.1.7 Sensible heat and latent from human body Q 4
Human heat output consists of sensible heat and latent heat Sensible heat is transferred from the body to the surrounding environment via convection, radiation, and conduction, while latent heat mainly comes from sweat evaporating from the skin surface and influencing the surrounding environment.
When calculating thermal loads for an adult, sensible heat and latent heat depend on the level of physical activity (which corresponds to the room's functional use) and the indoor temperature, while the total heat does not depend on room temperature The sensible heat, latent heat, and total heat values are obtained using specific formulas that relate activity level and temperature to heat outputs.
Sensible heat from human body Q4h:
Latent heat from human body Q4w:
Total heat gain from human body Q4:
Based on Table 6.2.2.1 chapter 6 [5] we have:
Table 2.10: Occupant density for each room function on level 2
Room/Zone Quantity Occupant density
Table 2.11: Sensible heat and latent from human Q4 on level 2
2.1.8 Sensible heat and latent from fresh air Q N
Air conditioning systems must continuously supply the conditioned space with a certain amount of fresh air to ensure adequate oxygen for human respiration When fresh air enters the room, it brings in heat, including sensible heat (QhN) and latent heat (QaN), which are quantified by the following expression.
Sensible heat from fresh air QhN:
Latent heat from fresh air QaN:
Total heat gain from human body Q4:
tN = 33.4 °C represents the incoming fresh air temperature, and tT = 24 °C is the temperature of the air after treatment before being supplied into the room dN denotes the humidity ratio of outdoor air, measured in g/kg.
Using tN = 33.4 °C and tT = 24 °C, the corresponding humidity ratios are dN = 26.463 g/kg and dT = 8.39 g/kg, as shown on the t-d chart The symbol l represents the required fresh-air rate per person per second (L/s), according to the ASHRAE Handbook – HVAC Applications (2011).
Tabel 2.12: Sensible heat and latent from fresh air QN on level 2
2.2.9 Sensible and latent heat from infiltration air Q 5
Air-conditioned rooms are typically sealed to preserve cooling, but human movement makes door opening and closing unavoidable, allowing outside air to infiltrate This air infiltration introduces heat transfer that disrupts the room’s temperature balance and reduces cooling efficiency, resulting in heat gain when outside air is warmer or heat loss when it is cooler The ensuing thermal fluctuations are irregular and do not follow a predictable pattern, complicating climate control.
Sensible heat from infiltration air Q5h:
Latent heat from infiltration air Q5a:
Outdoor air conditions for the ventilation system are defined by three key parameters: the incoming fresh air temperature (tN) at 33.4°C, the temperature of the air after treatment before it is supplied into the room (tT) at 24°C, and the humidity ratio of outdoor air (dN) measured in grams of water per kilogram of dry air (g/kg) These values guide HVAC design and indoor air quality management by specifying the temperature and moisture content of the air introduced to the indoor environment.
: The empirical factor is taken from Table 3.10 “Empirical Factors” For a room volume less than 500 m³, a value of 0.7 is selected.
Tabel 2.13: Sensible and latent heat from infiltration air Q5 of level 2
2.2.10 Heat loss from other sources Q 6
In addition to the heat sources mentioned above, there are several other heat gains that affect the cooling load, such as:
Sensible and latent heat emitted from heat exchangers, and hot or cold water pipes passing through the air-conditioned room.
Heat released by fans and heat loss through ductwork can warm the cooled air inside, but these losses are relatively small and can be neglected in the analysis Therefore, Q6 is set to zero.
2.2 Establish and calculate the air conditioning system diagram
2.2.1 Establish the air conditioning system diagram
Design and selection of equipment for the project system
3.1.1 Purpose of fresh air supply
An air-conditioned, closed environment can experience diminishing oxygen levels, which may cause occupants to feel short of breath or fatigued Supplying fresh air and improving ventilation are essential to compensate for the oxygen deficit and maintain comfortable, healthy indoor air quality in such spaces.
3.1.2 Operating principle of the fresh air supply system
In spaces with few occupants that are not fully sealed, using an exhaust fan is typically enough to remove stale air The resulting pressure differences draw fresh air in through door gaps, ensuring continuous ventilation.
In compact, enclosed spaces with high occupant density, when incoming air cannot meet demand, a dedicated fresh air supply fan must be used to ensure the building receives sufficient ventilation.
The fresh air supply fan helps accelerate the intake of outdoor air, balancing the natural environment with the air-conditioned environment.
3.1.3 Fresh air supply systems are currently being used
A fresh air supply system (supply-only) uses fans to push fresh outdoor air into a room, increasing indoor pressure As pressure builds, air naturally escapes through gaps around doors and windows, and when interior doors are opened, larger volumes can be expelled to maintain air balance and improve ventilation This approach delivers continuous fresh air while relying on natural exhaust paths to regulate pressure and circulate indoor air.
Fresh-air supply systems that operate exhaust-only use exhaust fans to draw air out of a room, lowering the internal pressure As the space becomes depressurized, fresh air is drawn in passively through door gaps or whenever doors are opened, helping to replace stale indoor air with outdoor air.
Two-way fresh air system using fans: Combines both supply and exhaust systems to balance indoor air pressure, allowing controlled intake and exhaust of air.
Two-way fresh air system with heat recovery: Provides ventilation while also helping to maintain a stable room temperature through heat exchange.
CALCULATION OF VENTILATION SYSTEM
Fresh air supply system
3.1.1 Purpose of fresh air supply
In an enclosed, air-conditioned space, oxygen levels can become depleted, leading to symptoms like shortness of breath and fatigue To protect occupant health and comfort, it is essential to introduce fresh air through effective ventilation that maintains adequate air exchange and offsets oxygen depletion.
3.1.2 Operating principle of the fresh air supply system
For sparsely occupied spaces that are not fully sealed, an exhaust fan can effectively remove stale air, while fresh air naturally flows in through door gaps driven by pressure differences between the indoors and outdoors This simple ventilation setup relies on these pressure differences to maintain air exchange without the need for additional intake equipment.
In small, enclosed spaces with high occupant density, incoming air often cannot meet the demand, making a fresh air supply fan essential to provide sufficient air for the building This dedicated ventilation solution augments airflow where natural ventilation falls short, helping to maintain indoor air quality and occupant comfort while enabling proper air changes per hour.
The fresh air supply fan helps accelerate the intake of outdoor air, balancing the natural environment with the air-conditioned environment.
3.1.3 Fresh air supply systems are currently being used
A supply-only fresh air system uses fans to push fresh air into a room, raising the indoor pressure The air then escapes naturally through gaps around doors and windows, or is expelled more quickly when doors are opened.
An exhaust-only fresh air supply system uses exhaust fans to remove air from the room, creating negative pressure that draws fresh outdoor air in through door gaps or whenever doors are opened This passive ventilation relies on leakage paths for air replacement, so the rate of air exchange depends on door gaps and the room’s airtightness, making it simple yet less controllable than balanced ventilation systems.
Two-way fresh air system using fans: Combines both supply and exhaust systems to balance indoor air pressure, allowing controlled intake and exhaust of air.
Two-way fresh air system with heat recovery: Provides ventilation while also helping to maintain a stable room temperature through heat exchange.
3.1.4 Determination of the speed at which air travels in the tube
Airspeed is a primary design parameter because it directly affects system performance When airspeed is high and fan capacity is large, noise levels increase, while the benefit of pipe size diminishes; conversely, lower airspeed with smaller fans reduces noise but constrains capacity Therefore, a suitable duct air velocity should be determined to keep the system operating stably while balancing noise with economic efficiency The centerline velocity in the duct is selected according to the orientations in Table 7.1 and Table 7.2 [1].
3.1.5 Calculating and verifying the airflow rate for the fresh air supply system
The airflow rate are determined by the following equation:
N: The number of human in the room
S: Area of the room (m 2 ) lN: Fresh air supply rate required per person (m3/h.person) lF: Fresh air supply rate required per square meter of building floor area (m3/h.m2) Determine the mass flow rate of supply air:
The volume flow rate of supply air is determined by:
GKL: Mass flow rate of supply air into the room (kg/s)
G: Volume flow rate of supply air into the room (m 3 /h)
QT: Sensible heat load of the supply air zone/space
(kW) hT – hH : Enthalpy difference between the room air and the supply air (kJ/kg)
The volume of fresh air supplied to the conditioned space is determined as follows:
Gfa: Fresh air supply flow rate to the conditioned space (m3/h) n: The number of human in the room l: Fresh air supply flow rate per person per second (l/s)
If the fresh air flow rate calculated from the above formula is less than 10% of the total supply air flow rate Gsa, the fresh air flow rate shall be set to 10% of Gsa, ensuring a minimum fresh-air contribution relative to the total supply air.
The return air flow rate:
The fresh air is supplied directly from the outside into the evaporators in the room through the airducts.
Determine the airduct section by the following equation:
G: The air flowrate into the airduct (m 3 /h) v: Velocity of the air inside airduct (m/s)
Based on the calculated duct cross-sectional area, select an actual duct size with a larger and closest available cross-sectional area (based on table 7.3 [1]).
3.1.7 Determine actual air velocity in the duct
In which: vtt S tt ( m⁄s) (3.8) vtt: Actual air velocity in the duct (m/s)
Stt: Actual cross-sectional area of the duct segment (m²)
3.1.8 Use Duct Checker Pro to determine air duct dimensons
Duct Checker is a specialized software tool for engineers and HVAC professionals to design and verify air duct systems It enables users to calculate air velocity, pressure loss, and duct sizing with ease and accuracy By streamlining complex calculations and providing a user-friendly input interface, Duct Checker helps optimize duct performance, improve system efficiency, and ensure compliance with industry standards such as ASHRAE Suitable for both small-scale projects and large commercial installations, it provides a reliable solution for efficient HVAC ductwork planning and analysis.
Typical calculation for the fresh air duct shaft from the PAU.AT.02 – Level 2
Fig 3.2: Duct shaft PAU.AT.02 – level 2
First, click to setting before strating calculate duct size
Fig 3.3: Setting table of Duct Checker Pro
After checking and setting values, we can start calculate duct size:
First step: Return to calculate page and enter the flow rate.
Second step: Click to the calculator to calculate duct size After that, on the screen, there will be many types of air duct sizes.
Third step: Choose the appropriate size.
Example: The flow rate through AB duct is 1000 m 3 /h We choose the duct size 350x200 for the main duct with velocity of 3.97 m/s and pressure loss of 0.819 Pa/m
Fig 3.4: Calulate screen of Duct Checker Pro
Below is a summary table of the parameters of duct size on duct shaft PAU.AT.02 on second floor.
Table 3.1: The parameters of duct size on duct shaft PAU.AT.02 – level 2.
Conclusion: The duct sizing process revealed discrepancies with the design company’s specifications, driven by objective factors such as aesthetics, available space, and the need for uniformity across floors In alignment with the priority group, selecting ducts that maintain an approximately 1 Pa/m pressure loss is recommended to optimize airflow and overall system performance.
3.1.9 Determination of flexible air duct size
Flexiable air duct size is calculated by the following equation:
Q: Airflow through the duct (m3/s) v: Wind velocity (m/s) Select v = 3 ÷ 3.5 m/s.
Example: Calculate the flexible duct of branch duct on duct shaft PAU.AT.02 – level 2 With a fresh air supplied flow rate is Q = 100 m 3 /h = 0.027 m 3 /s, we get:
4Q d =√πv =√4 × 0.027 π × 3.5 = 0.110 m Thus, with d = 0.110 m, we choose the flexible duct size of φ150mm
3.1.10 Determine pressure loss in air duct
The pressure loss in the air duct is divided into two parts:
∆Pₘₛ: Pressure loss due to friction along the duct (Pa)
∆Pcb: Local pressure loss at duct fittings (tees, elbows, take-offs, etc.) (Pa) a) Friction loss
The frictional resistance of the air duct segment is determined using the following formula:
With: ΔPms: Total friction loss along the duct segment (Pa) l: Length of the duct segment considered for pressure loss (m) ΔPl: Friction loss per meter of duct (Pa/m)
Friction loss per meter in a duct (ΔPl) is determined by the duct’s equivalent diameter and the actual air velocity, but to simplify calculations, a uniform friction loss method is used, applying the same loss per meter along the entire duct length (typically ΔP1 = 1 Pa/m).
The total length of the fresh air supply duct from PAU.AT.02 – level 2 (considering only the segment with the highest loss usually the furthest segment) is 26m.
Therefore, the total friction loss in the fresh air supply duct shaft is: ΔPms = l ΔPl = 26 × 1 = 26 (Pa) b) Local pressure loss
For local pressure loss at duct fittings (nodes), it is considered that ΔPcb = ΔPms. Accordingly, the formula for calculating the local pressure loss is:
With: ΔPcb: Local pressure loss through circular or rectangular elbows (Pa) ltd = a × d: Equivalent length of the fittings, based on Table 7.4 [1]
Local pressure loss at tees, branches, transitions, and enlargements
The local pressure loss through tees or branches is determined according to [1] using the following formula:
With: ΔPcb: Local resistance of tees, branches, transitions, and enlargements (Pa) n: Dynamic pressure coefficient, obtained from Tables 7.7 to 7.10 [1] pd: Dynamic pressure (Pa), determined using Table 7.6 [1]
Although dynamic pressure (pd) is normally derived from Table 7.6, the project involves a large number of fittings contributing to local pressure losses, and several components listed in the textbook have limitations regarding airflow direction and fitting geometry To streamline the calculation, the team will determine the local pressure loss of these components using the ASHRAE Duct Fitting Database software.
Use ASHRAE Duct Fitting Database to check local pressure loss:
ASHRAE's Duct Fitting Database is a comprehensive, authoritative resource created by the American Society of Heating, Refrigerating and Air-Conditioning Engineers to help HVAC professionals accurately estimate pressure losses in duct systems It catalogs hundreds of common air-distribution fittings—elbows, transitions, branches, and take-offs—with loss coefficients derived from rigorous laboratory testing and research, enabling precise design and analysis calculations Widely integrated with HVAC design software and engineering practice, this database supports optimized energy efficiency and dependable airflow performance across mechanical systems.
Fig 3.5: ASHRAE Duct Fitting Database
Typical checking local pressure loss for the fresh air duct shaft from the PAU.AT.02 – Level 2:
Example: On CG duct we have 2 90 o elbows with dimension of 250x200 and flow rate of 700 m 3 /h According to ASHRAE sofware, the pressure loss from 90 o elbow is
Pressure loss from fire damper: According to ASHRAE sofware, the pressure loss from fire damper is 2 Pa on CG duct and 3 Pa on HK duct.
Fig 3.6: Pressure loss from fire damper of fresh air duct
Pressure loss from VCD: Based on the result of ASHRAE sofware, the pressure loss from VCD is 0 Pa So we can ignore this pressure loss.
Fig 3.7: Pressure loss from VCD of fresh air duct
Pressure loss from diverging and transition:
An example for an HJ duct analyzes the diverging section and transition at a flow rate of 300 m³/h ASHRAE software results show a pressure loss of 4 Pa across the diverging and transition sections.
Fig 3.8: Pressure loss from diverging and transition of fresh air duct
Following the determination of all pressure losses in the air duct, the summary table for the pressure loss in the fresh air duct shaft (PAU.AT.02 – Level 2) is presented, as shown in Table 3.2.
Table 3.2: Total local pressure loss of fresh air duct shaft from the PAU.AT.02 – Level 2
Detail pressure loss Total local pressure loss (Pa)
Thus, we have the total pressure loss of fresh air duct shaft from the PAU.AT.02 –
Exhaust air system
Exhaust ventilation systems in enclosed spaces such as toilets, parking garages, and storage rooms are essential for maintaining indoor air quality, protecting human health, and supporting the safety and hygiene of the built environment Each functional area generates distinct pollutants and associated ventilation needs, from moisture and odors in restrooms to vehicle exhaust and chemical fumes in parking facilities, and dust or volatile compounds in storage rooms Proper design, installation, and operation of these systems help control contaminants, reduce exposure, and enhance occupant comfort Tailored ventilation strategies—encompassing airflow rates, filtration, exhaust placement, and maintenance schedules—address the specific risks of each space while aligning with applicable standards and codes Integrating exhaust ventilation into building design promotes healthier indoor environments and minimizes pollutant buildup, odors, and moisture-related damage.
Toilets are enclosed spaces where odours, high humidity, and harmful gases such as ammonia (NH3) and hydrogen sulfide (H2S) can accumulate Without proper ventilation, these conditions cause discomfort, promote the growth of mould, bacteria, and other pathogens, and degrade indoor air quality An efficient exhaust system continuously removes polluted air, improving air freshness, reducing health risks, and maintaining hygiene standards in the facility.
Underground and enclosed parking garages are major sources of air pollution because vehicle emissions introduce gases such as carbon monoxide (CO), nitrogen oxides (NOx), and volatile organic compounds (VOCs) into the indoor environment If these pollutants are not effectively removed, they can reach dangerous levels, posing health risks to occupants and maintenance personnel Exhaust ventilation systems are essential to dilute and exhaust contaminated air, ensuring compliance with indoor air quality standards and occupational health regulations and preventing long-term exposure for those inside the facility.
Storage areas such as warehouses and storage rooms can release fumes, odors, and gases from chemicals, flammable materials, or organic substances, particularly when ventilation is inadequate Stagnant air, dust, and accumulated fumes raise the risk of explosions, corrosion, and product damage Implementing proper ventilation helps regulate temperature, humidity, and contaminant levels, supporting fire safety, protecting equipment, and creating a healthier work environment.
3.2.2 Calculation of exhaust air flow rate
Exhuast air flow rate is detemined by the following equation:
V: Volume of the space requiring ventilation (m³)
ACH: Number of air changes per hour, referenced from Appendix G [2]
Typical calcution for basement B2 toilet
Table 3.3: Air exchange flow rate of basement B2 toilet
Room/Zone Volume (m 3 ) Air change rate Flow rate (m 3 /h)
3.2.3 Calculation of dimensions of exhuast air duct
Similar to fresh air supply duct system, we determine dimensions of exhaust air duct by Duct Checker Pro software.
Typical ditermination of exhuast air duct for basement B2 toilet
Table 3.4: Parameters of exhuast air duct for basement B2 toilet
3.2.4 Pressure loss through exhaust air duct system
Typical checking for exhuast air duct for basement B2 toilet
Friction pressure loss: The total length of exhuast air duct for basement B2 toilet is 22m Therefore, the total friction loss in the exhuast air duct system is: ΔPms = l ΔPl = 22 × 1 = 22 (Pa)
Local pressure loss: Similar to fresh air supply duct system, we check pressure loss through of exhaust air duct by ASHRAE Duct Fitting Database.
Table 3.5: Local pressure loss of exhaust air duct of basement B2 toilet
Detail pressure loss Total local pressure loss (Pa)
Thus, total pressure loss of exhaust air duct of basement B2 toilet is: ΔP = ΔPms + ΔPcb = 22 + 84 = 106 (Pa)
3.2.5 Choose fan for exhaust air duct system
To choose a fan for exhaust air duct system, we use Fantech sofware.
With a total airflow of 2,790 m3/h and a total pressure loss of 106 Pa, the system meets the required exhaust performance After searching with Fantech for an exhaust air duct fan for the basement B2 toilet, the suitable model is the SCEEC35 – Short Case EC Series – Axial Fans, with its parameters shown in Fig 3.10.
Fig 3.10: Model fan SCEEC35 – Short Case EC Series – Axial Fans
CALCULTIONS AND CHECKING RESULTS
Results of excess heat calculation using the Carrier method
4.1.1 Results of sensible radiant heat gain through glass Q 11
Table 4.1: Results of sensible radiant heat gain through glass Q11
Room/Zone Direction Glass area
4.1.2 Results of sensible heat gain through walls Q 22
Table 4.2: Results of sensible heat gain through walls Q22
4.1.3 Results of sensible heat gain through floor Q 23
Table 4.3: Results of sensible heat gain through floor Q23 of project
4.1.4 Results of sensible heat gain from lighting Q 31
Table 4.4: Results of sensible heat gain from lighting Q31
4.1.5 Results of sensible heat gain from equipment Q 32
Table 4.5: Results of sensible heat gain from equipment Q32
4.1.6 Results of sensible heat and latent from human body Q 4
Table 4.6: Results of sensible heat and latent from human body Q4
4.1.7 Results of sensible heat and latent from fresh air Q N
Table 4.7: Results of sensible heat and latent from fresh air QN
4.1.8 Results of sensible and latent heat from infiltration air Q5
Table 4.8: Results of sensible and latent heat from infiltration air Q5
Comparision of cooling load of the project
Table 4.9: Comparision of cooling load of the project
The error between Design and Carrier method (%)
The error between Design and DK BIM (%)