The Bernoulli equation can be developed by equating the forces on an element of a stream tube in a frictionless fluid flow to the rate of momentum change. On integrating this relationship for steady flow, the following expression (Osborne 1966) results:
where
| v = streamline (local) velocity, m/s |
|
| p = absolute pressure, Pa (N/m2) |
| ρ = density, kg/m3 |
| g = acceleration caused by gravity, m/s2 |
| z = elevation, m |
Assuming constant fluid density in the system, Equation (1) reduces to
Although Equation (2) was derived for steady, ideal frictionless flow along a stream tube, it can be extended to analyze flow through ducts in real systems with friction and dynamic losses. In terms of pressure, the relationship for fluid resistance between two sections is
where
| V = average duct velocity, m/s |
| Δpt,1–2 = total pressure loss caused by friction and dynamic losses between sections 1 and 2, Pa |
In Equation (3), V (cross-section average velocity) replaces v (streamline velocity) because experimentally determined loss coefficients allow for errors in calculating v 2/2gc (velocity pressure) across streamlines.
On the left side of Equation (3), add and subtract pz1; on the right side, add and subtract pz2, where pz1 and pz2 are the values of atmospheric air pressure at heights z1 andz2. Thus,
Atmospheric air pressure at any elevation ( pz1 and pz2) expressed in terms of the atmospheric pressure pa at the same datum elevation is given by
Substituting Equations (5) and (6) into Equation (4) and simplifying yields the total pressure change between sections 1 and 2. Assume no temperature change between sections 1 and 2 (e.g., no heat exchanger or significant heat loss or gain within the section); therefore, ρ1 = ρ2. When a heat exchanger is located in the section, the arithmetic average of the heat exchanger inlet and outlet temperatures is generally used (with the heat exchanger treated as a duct section). Let ρ = ρ1 = ρ2, and ps,1 = ( p1 − pz1) and ps,2 = ( p2 − pz2) are gage pressures at elevations z1 and z2.
Rearranging Equation (7b) yields
where
| ps,1 = static pressure, gage at elevation z1, Pa |
| ps,2 = static pressure, gage at elevation z2, Pa |
| V1 = average velocity at section 1, m /s |
| V2 = average velocity at section 2, m /s |
| ρa = density of ambient air, kg /m3 |
| ρ = density of air or gas in duct, kg/m3 |
| Δ pse = thermal gravity effect, Pa |
| Δ pt = total pressure change between sections 1 and 2, Pa |
| Δ pt,1–2 = total pressure loss caused by friction and dynamic losses plus thermal gravity effects between sections 1 and 2, Pa |
The terms head and pressure are often used interchangeably; however, head is the height of a fluid column supported by fluid flow, whereas pressure is the normal force per unit area. For liquids, it is convenient to measure head in terms of the flowing fluid. With a gas or air, however, it is customary to measure pressure exerted by the gas on a column of liquid.
The term ps/ρg is static head; ps is static pressure.
The term V 2/2g refers to velocity head, and ρV2/2 refers to velocity pressure. Although velocity head is independent of fluid density, velocity pressure [Equation (8)] is not.
where
| pv = velocity pressure, Pa |
| V = fluid mean velocity, m /s |
|
For air at standard conditions (1.204 kg/m3), Equation (8) becomes
Velocity is calculated by
where
| Q = airflow rate, L /s |
| A = cross-sectional area of duct, m2 |
Total pressure is the sum of static pressure and velocity pressure:
or
where
| pt = total pressure, Pa |
| ps = static pressure, Pa |
Pressure and Velocity Measurements
ASHRAE Standard 41.2-2018 provides procedures for measuring air pressure and velocity. It also describes the range, precision, and limitations of instruments used for these purposes. A manometer is a simple and useful means for measuring partial vacuum and low pressure. Static, velocity, and total pressures in a duct system relative to surrounding space pressures can be measured with a pitot tube connected to a manometer. Pitot tube construction and locations for traversing round and rectangular ducts are also presented in Standard 41.2-2018.
The total pressure change caused by friction, fittings, equipment, and net thermal gravity effect for each section of a duct system is calculated by the following equation:
where
| Δpti = net total pressure change for i sections, Pa |
| Δpfi = pressure loss caused by friction for i sections, Pa |
| Δpij = total pressure loss caused by j fittings, including fan system effect (FSE), for i sections, Pa |
| Δ pik = pressure loss caused by k equipment for i sections, Pa |
| Δpseir = thermal gravity effect caused by r stacks for i sections, Pa |
| m = number of fittings within i sections |
| n = number of equipment within i sections |
| λ = number of stacks within i sections |
| nup = number of duct sections upstream of fan (exhaust/return air subsystems) |
| ndn = number of duct sections downstream of fan (supply air subsystems) |
From Equation (7), the thermal gravity effect for each nonhorizontal duct with a density other than that of ambient air is determined by the following equation:
where
| Δpse = thermal gravity effect, Pa |
| z1 and z2 = elevation from datum in direction of airflow (Figure 1), m |
| ρa = density of ambient air, kg/m3 |
| ρ = density of air or gas within duct, kg / m3 |
| |
| g = 9.81 = gravitational acceleration, m /s2 |
Example 1.
For Figure 1, calculate the thermal gravity effect for two cases: (a) air cooled to −34°C, and (b) air heated to 540°C. Density of air at −34°C is 1.477 kg /m3 and at 540°C is 0.434 kg /m3. Density of ambient air is 1.204 kg /m3. Stack height z = z2 − z1 is 15 m.
Solution:
(a) For ρ > ρa (Figure 1A),
(b) For ρ < ρa (Figure 1B),
Example 2.
Calculate the thermal gravity effect for the two-stack system shown in Figure 2, where the air is 120°C and stack heights are 15 and 30 m. Density of 120°C air is 0.898 kg /m3; ambient air is 1.204 kg /m3.
Solution:
To determine the fan total pressure requirement for a system, use the following equation:
where
| Fup and Fdn = sets of duct sections upstream and downstream of fan |
| Pt = fan total pressure, Pa |
| ε = symbol that ties duct sections into system paths from exhaust/return air terminals to supply terminals |
Figure 3 shows the use of Equation (15). This system has three supply and two return terminals consisting of nine sections connected in six paths: 1-3-4-9-7-5, 1-3-4-9-7-6, 1-3-4-9-8, 2-4-9-7-5, 2-4-9-7-6, and 2-4-9-8. Sections 1 and 3 are unequal area; thus, they are assigned separate numbers in accordance with the rules for identifying sections (see step 5 in the section on HVAC Duct Design Procedures). To determine the fan pressure requirement, apply the following six equations, derived from Equation (15). These equations must be satisfied to attain pressure balancing for design airflow. Relying entirely on dampers is not economical and may create objectionable flow-generated noise.
Example 3.
For Figures 4A and 4C, calculate the thermal gravity effect and fan total pressure required when the air is cooled to −30°F. The heat exchanger and ductwork (section 3 to 2) total pressure losses are 0.70 and 0.28 in. of water respectively. The density of −34°C air is 1.477 kg/m3; ambient air is 1.204 kg/m3. Elevations are 21 and 3 m.
Solution:
(a) For Figure 4A (downward flow),
(b) For Figure 4C (upward flow),
Example 4.
For Figures 4B and 4D, calculate the thermal gravity effect and fan total pressure required when air is heated to 120°C. Heat exchanger and ductwork (section 3 to 2) total pressure losses are 170 and 70 Pa, respectively. Density of 120°C air is 0.898 kg/m3; ambient air is 1.204 kg/m3. Elevations are 21 and 3 m.
Solution:
(a) For Figure 4B (downward flow),
(b) For Figure 4D (upward flow),
Example 5.
Calculate the thermal gravity effect for each section of the system in Figure 5 and the system’s net thermal gravity effect. Density of ambient air is 1.204 kg/m3, and the lengths are as follows: z1 = 15 m, z2 = 27 m, z4 = 30 m, z5 = 8 m, and z9 = 60 m. Pressure required at section 3 is −25 Pa. Write the equation to determine the fan total pressure requirement.
Solution: The following table summarizes the thermal gravity effect for each section of the system as calculated by Equation (14). The net thermal gravity effect for the system is 118 Pa. To select a fan, use the following equation:
|
Path (x – x′) |
Temp., °C |
ρ, kg/m3 |
Δz (zx′ − zx), m |
Δρ (ρa − ρx–x′), kg/m3 |
Δpse , Pa [Eq. (14)] |
|---|
|
1–2 |
815 |
0.324 |
(27 − 15) |
+0.880 | | |
+104 | | |
|
3–4 |
540 |
0.434 |
0 |
+0.770 | | |
0 | |
|
4–5 |
540 |
0.434 |
(8 − 30) |
+0.770 | | |
−166 | |
|
6–7 |
120 |
0.898 |
0 |
+0.306 | | |
0 | |
|
8–9 |
120 |
0.898 |
(60 − 0) |
+0.306 | | |
+180 | |
|
Net Thermal Gravity Effect |
118 | |
Airflow rate and direction depend on the sum of the stack, wind, and mechanically induced driving forces. Chapter 16 provides guidance on wind effects and on combining all of these driving forces.
2.1 PRESSURE CHANGES IN SYSTEM
Figure 6 shows total and static pressure changes in a fan/duct system consisting of a fan with both supply and return air ductwork. Also shown are total and static pressure gradients referenced to atmospheric pressure.
For all constant-area sections, total and static pressure losses are equal. At diverging transitions, velocity pressure decreases, absolute total pressure decreases, and absolute static pressure can increase. The static pressure increase at these sections is known as static regain.
At converging transitions, velocity pressure increases in the direction of airflow, and absolute total and absolute static pressures decrease.
At the exit, total pressure loss depends on the shape of the fitting and the flow characteristics. Exit loss coefficients Co can be greater than, less than, or equal to one. Total and static pressure grade lines for the various coefficients are shown in Figure 6. Note that, for a loss coefficient less than one, static pressure upstream of the exit is less than atmospheric pressure (negative). Static pressure just upstream of the discharge fitting can be calculated by subtracting the upstream velocity pressure from the upstream total pressure.
At section 3, total pressure loss depends on the shape of the entry. Total pressure immediately downstream of the entrance equals the difference between the upstream pressure, which is zero (atmospheric pressure), and loss through the fitting. Static pressure of ambient air is zero; several diameters downstream, static pressure is negative, equal to the sum of the total pressure (negative) and the velocity pressure (always positive).
System resistance to airflow is noted by the total pressure grade line in Figure 6. Sections 3 and 4 include fan system effect pressure losses. To obtain the fan static pressure rise requirement for selecting fans rated using this parameter (i.e., for fans with unducted outlets, such as plenum fans) and where fan total pressure rise is known, use
where
| Ps = fan static pressure rise, Pa |
| Pt = fan total pressure rise, Pa |
| pv,o = fan outlet velocity pressure, Pa |
Fan static pressure rise is also the difference between the fan outlet static pressure and the fan inlet total pressure (ps,o − pt,i). However, note that fan static pressure rise is not the difference in static pressure between the fan inlet and fan outlet, nor is it the sum of static pressure differences for components in the air-handling system.
6.1 DESIGN CONSIDERATIONS
See Chapter 19 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment for (1) sealant specifications, and (2) the rationale for HVAC system sealing and leakage testing.
System Sealing. All ductwork and plenum transverse joints, longitudinal seams, and duct penetrations should be sealed. Longitudinal seams are joints in the direction of airflow. Transverse joints are connections of two duct sections, with the connections oriented perpendicular to airflow. Openings for rotating shafts, wires, and pipes or tubes should be sealed with bushings or other devices that minimize air leakage but that do not interfere with shaft rotation or prevent thermal expansion. Spiral lock seams need not be sealed. Duct-mounted equipment, such as terminal units, reheat coils, access doors, sound attenuators, balancing dampers, control dampers, and fire dampers, should be specified as low leakage so that the system can meet the air leakage acceptance criteria set by the designer, standards, and codes. Recommended specifications for duct-mounted equipment are provided later in this section.
Sealing that would void product listings (e.g., for fire/smoke or volume control dampers) is not required. However, low-leakage duct-mounted components, including terminal units, reheat coils, and access doors, should be specified so that the combined HVAC system air leakage will not exceed criteria set by the designer, ASHRAE Handbook, standards, and codes. For example, some UL-listed and UL-labeled fire/smoke dampers allow sealing and gasketing of breakaway duct/sleeve connections; all can provide sealed non-breakaway duct/sleeve connections.
Scope. It is recommended that supply air (both upstream and downstream of the VAV box primary air inlet damper when used) and independent exhaust air systems or selected parts thereof be tested for air leakage after construction at operating conditions using ASHRAE Standard 215 (ASHRAE 2018) to verify (1) good workmanship, and (2) the use of low-leakage components as required to achieve the design allowable system air leakage. Testing these particular systems is important because of the potentially significant energy impacts caused by air leakage to or from them. As a minimum, 25% of the system based on duct surface area should be tested, and an additional 25% should be tested if any of the initial sections fail. If any section of the second 25% fails, the entire system should be tested. Sections to be tested should be selected randomly by the owner’s representative. These leakage tests should be conducted by an independent party responsible to the owner’s representative after the system sections to be tested are fully assembled but before the installation of insulation and concealment of ductwork. To ensure that a system passes its air leakage test at operating conditions, sufficient ductwork sections should be leak tested by pressurization during construction to vet construction techniques. To supplement these construction phase tests, leakage of duct-mounted components should be determined by specification and certified leakage data provided with equipment submittals. Leakage of air-handling units should be determined by specification and verification leakage tests.
|
Independent exhaust air system: air discharged from a space to the outdoors by a system not coupled to supply or return air systems. |
Acceptance Criteria. To enable proper accounting of leakage-related impacts on fan energy and space conditioning loads, the allowable system air leakage for each fan system test section should be established by the design engineer as a percentage of fan airflow entering the test section at a reference operating condition. The method of test in ASHRAE Standard 215-2018 includes how to establish a reference operating condition, but because it is not a rating standard, it does not provide system air leakage acceptance criteria. However, as described in Chapter 19 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment, the recommended maximum system leakage is 5% of design airflow. Exceptions: supply and return ductwork sections that leak directly to/from outdoors and exhaust system ductwork sections that draw in indoor air through leaks should be limited to 2%.
Equation (37) is for use in a leakage test during construction to ensure that ductwork will meet the leakage specification. This equation translates system fractional air leakage to test section leakage class, as specified by ASHRAE Standard 90.1-2019, section 6.4.4.2.2.
where
| CL,section = test section leakage class, L/s per Pa0.65 per square metre of duct surface area |
| Qfan = maximum fan airflow that would occur during operation, L/s} |
| Qleak, frac = system leakage fraction corresponding to maximum fan airflow that would occur during operation, % |
| Asystem = total system duct surface area, m2 |
| Δ psection = test section static pressure difference corresponding to maximum fan airflow that would occur during operation, Pa |
Equation (37) shows that leakage class depends on fractional air leakage and normalized fan airflow (Qfan/Asystem), and varies inversely with the pressure difference raised to the 0.65 power. For example, to achieve 3% leakage for ductwork with Qfan/Asystem = 10 L/s per m2 of duct surface area and Δ psection = 750 Pa, the required leakage class is 0.004 L/s per Pa0.65 per square metre of duct surface area. With Δ psection = 125 Pa instead (six times less), the required leakage class is about three times greater (0.13).
The maximum acceptable air leakage for a test section corresponding to the leakage class determined by Equation (37) can be expressed by Equation (38).
wherewhere
| Qleak,section = test section air leakage, L/s |
| Asection = test section duct surface area, m2 |
Equations (37) and (38) can be combined so that the maximum acceptable air leakage for a test section during construction is simply a function of system fractional leakage, normalized fan airflow (Qfan/Asystem), and section duct surface area:
Thus, for air leakage tests during construction, the maximum acceptable leakage for a ductwork section is given by Equation (38) or (39). Duct surface area should be calculated in accordance with European Standard EN 14239. The test pressure for each section should be specified by the design engineer based on the maximum static pressure for that section that would occur during operation at maximum fan airflow.
Example 7.
The system depicted by Figure 17 has the characteristics summarized by Table 8. For a maximum acceptable leakage of 3% of the 2360 L/s maximum fan airflow, which is 71 L/s total, what is the maximum allowable ductwork leakage in (1) each section that is to be tested at the pressures noted, and (2) Sections 4 and 5 when leak tested together?
Solution: The calculations and maximum allowable leakage are also summarized in Table 8. Note that adjacent sections with the same static pressure can be grouped for leakage testing. In this case, Sections 4 and 5 can be grouped. The allowable leakage for Sections 4 and 5 when tested individually is 4 L/s and 20 L/s respectively. The allowable leakage for Sections 4 and 5 when tested together is 24 L/s.
Recommended Specification for Duct-Mounted Equipment. Duct-mounted component leakage is controlled by specification and certified leakage data provided with equipment submittals. The following are recommended leakage-related specifications for duct-mounted equipment.
Terminal Units. Seal longitudinal seams of casings, inlet face of casings, and inlet collars with mastic. Seal damper shaft penetrations of casings. Casing leakage for the basic terminal unit should not exceed 2.1 L/s at 249 Pa static pressure differential. Terminal unit leakage with an access door should not exceed 2.2 L/s at 249 Pa. Testing should be by an accredited laboratory. Leakage tests should comply with ASHRAE Standard 130.
Access doors in terminal unit casings should comply with AMCA Standard 500-D, and leakage rates should be certified per AMCA (2013) Publication 511. Access door (frame not included) leakage should not exceed 0.24 L/s at 2.49 kPa static pressure differential.
Electric Reheat Coils. Flange-mount electric coils with a gasket. Coil leakage should not exceed 0.24 L/s at 249 Pa static pressure differential. Tests should be by an accredited laboratory. Leakage tests should be in compliance with ASHRAE Standard 126-2016 (ASHRAE 2016).
Hot-Water Reheat Coils. Flange-mount hot-water coils with a gasket in an insulated plenum or casing. Seal all seams and casing penetrations for supply and return water tubes. Coil casing leakage (not counting transverse joints) should not exceed 0.24 L/s at 249 Pa static pressure differential. An accredited laboratory should perform leakage tests in compliance with ASHRAE Standard 126.
Access Doors. Access door leakage (excluding the frame) should not exceed 0.24 L/s at 2.49 kPa static pressure differential. Leakage tests should comply with AMCA Standard 500-D, and leakage rates certified per AMCA (2013) Publication 511.
Attenuators. Sound attenuator casing seams should be sealed at the factory if used in-line with the ductwork. Attenuators stacked in plenums do not have to be sealed, but plenums should be sealed.
Fire/Smoke Dampers. These dampers should be installed in accordance with the manufacturer’s UL installation instructions. Each fire damper should be furnished with a UL-approved sleeve. Sleeve seams should be continuously welded or sealed, and the transverse joint should be a sealed UL-approved flanged duct sleeve connection (break-away or non-break-away).
Balancing Dampers. Balancing damper casing seams should be continuously welded or sealed, and the shaft penetrating the casing should have seals.
Control Dampers. Control damper shafts penetrating ducts should have seals.
Recommended Specification for Air-Handling Units. Refer to Chapter 19 of the 2020 ASHRAE Handbook—HVAC Systems and Equipment for leakage-related specifications for air-handling units.
Responsibilities. The engineer should
Specify HVAC system components, duct-mounted equipment, sealants, and sealing procedures that together will meet the system airtightness design objective.
Inspect the system during construction for quality of workmanship and to verify that correct duct-mounted components and air-handling units are installed.
Specify the construction-stage ductwork leakage test standard or procedures.
Specify the construction-stage test pressures to the nearest 25 Pa expected during operation at design conditions and the maximum allowable air leakage.
Review and approve the sheet metal and test contractors’ leakage test reports. If any system has a leakage failure, the engineer should discuss remedies with the sheet metal contractor, vendor, and/or owner’s representative.
The sheet metal contractor should
Construct the system using quality workmanship and correct duct-mounted components and air-handling units. If any installed duct-mounted equipment appears to be leakage suspect, the contractor should discuss remedies with the engineer and/or owner’s representative.
Conduct ductwork leakage pressurization tests during construction. As a minimum, 25% of the ductwork system (based on duct surface area) should be tested during construction, and another 25% if any of the initial sections fail. If any section of the second 25% fails, the entire ductwork system should be leak tested.
Provide connections for test apparatus, and separate test sections from each other as needed so that the test apparatus capacity is not exceeded.
Report test results and, where required, take corrective action to seal ductwork and absorb the cost for conducting related additional leak tests.
The test contractor should
Conduct the operating system leakage test in compliance with ASHRAE Standard 215-2018.
Report test results, including reasons for any failures.
It should be the responsibility of the owner or owner’s representative to provide direction upon request.
Because duct systems can convey smoke, hot gases, and fire from one area to another and can accelerate a fire within the system, fire protection is an essential part of air-conditioning and ventilation system design. Generally, fire safety codes require compliance with the standards of national organizations. NFPA Standard 90A examines fire safety requirements for (1) ducts, connectors, and appurtenances; (2) plenums and corridors; (3) air outlets, air inlets, and fresh air intakes; (4) air filters; (5) fans; (6) electric wiring and equipment; (7) air-cooling and -heating equipment; (8) building construction, including protection of penetrations; and (9) controls, including smoke control.
Fire safety codes often refer to the testing and labeling practices of nationally recognized laboratories, such as Factory Mutual and Underwriters Laboratories (UL). UL’s annual Building Materials Directory lists fire and smoke dampers that have been tested and meet the requirements of UL Standards 555 and 555S. This directory also summarizes maximum allowable sizes for individual dampers and assemblies of these dampers. Fire dampers are 1.5 h or 3 h fire-rated. Smoke dampers are classified by (1) temperature degradation [ambient air or high temperature (120°C minimum)] and (2) leakage at 250 and 1000 Pa pressure difference (2 and 3 kPa classification optional). Smoke dampers are tested under conditions of maximum airflow. UL’s annual Fire Resistance Directory lists fire resistances of floor/roof and ceiling assemblies with and without ceiling fire dampers.
For a more detailed presentation of fire protection, see the NFPA (2008) Fire Protection Handbook, Chapter 54 of the 2019 ASHRAE Handbook—HVAC Applications, and Klote et al. (2012).
In all new construction (except low-rise residential buildings), air-handling ducts and plenums that are part of an HVAC air distribution system should be thermally insulated in accordance with ASHRAE Standard 90.1. Duct insulation for new low-rise residential buildings should comply with ASHRAE Standard 90.2. Existing buildings should meet requirements of ASHRAE Standard 100. In all cases, thermal insulation should meet local code requirements. Insulation thicknesses in these standards are minimum values; economic and thermal considerations may justify higher insulation levels. Additional insulation, vapor retarders, or both may be required to limit vapor transmission and condensation.
Duct heat gains or losses must be known to calculate supply air quantities, supply air temperatures, and coil loads. To estimate duct heat transfer and entering or leaving air temperatures, refer to Chapters 4 and 23.
Ducts entering into spaces considered secured or sensitive should contain measures to detect or inhibit forced entry into those spaces through the duct system. Inhibiting measures should be based on a delay or resistance time set by the facility owner or user. The following security measures should be considered:
Barrier duct bars: placement should be at locations deemed appropriate by the facility owner and user, or where secured or sensitive area boundaries exist (e.g., duct penetration through the wall of a secured room). Bar diameter, material, and spacing should be established by the facility owner or user. Ensure that barriers do not inhibit proper operation and maintenance of dampers, detectors, sensors, and other devices in the duct system.
Welded diffuser and return grilles: grille material and spacing should be established by the facility owner or user.
Sensors on access doors: sensors should be connected to a facility security system or panel, and notify security personnel of a duct system breach. Communicating HVAC maintenance schedules to security personnel is necessary to ensure that an intrusive duct system breach is not confused with HVAC system maintenance.
Use Figure 18 for preliminary sizing of air intake and exhaust louvers. For air quantities greater than 3300 L/s per louver, the air intake gross louver openings are based on 2 m /s; for exhaust louvers, 2.5 m /s is used for air quantities of 2400 L/s per louver and greater. For smaller air quantities, see Figure 18. These criteria are presented on a per-louver basis (i.e., each louver in a bank of louvers) to include each louver frame. Representative production-run louvers were used in establishing Figure 18, and all data used were based on AMCA Standard 500-L tests. For louvers larger than 1.5 m2, the free areas are greater than 45%; for louvers less than 1.5 m2, free areas are less than 45%. Unless specific louver data are analyzed, no louver should have a face area less than 0.4 m2. If debris can collect on the screen of an intake louver, or if louvers are located at grade with adjacent pedestrian traffic, louver face velocity should not exceed 0.5 m /s.
Louvers require special treatment because the blade shapes, angles, and spacing cause significant variations in louver-free area and performance (pressure drop and water penetration). Selection and analysis should be based on test data obtained from the manufacturer in accordance with AMCA Standard 500-L, which presents both pressure drop and water penetration test procedures and a uniform method for calculating the free area of a louver. Tests are conducted on a 1220 mm square louver with the frame mounted flush in the wall. For water penetration tests, rainfall is 100 mm / h, no wind, and the water flow down the wall is 0.05 L/s per linear metre of louver width.
AMCA Standard 500-L also includes a method for measuring water rejection performance of louvers. These louvers are subjected to simulated rain and wind pressure and tested at a rainfall of 76 mm / h falling on the louver’s face with a predetermined wind velocity directed at the face of the louver (typically 13 or 20 m /s). Effectiveness ratings are assigned at various airflow rates through the louver.
No Space Constraints. Round ductwork is preferable to rectangular or flat oval ductwork when adequate space is available for the following reasons.
Mass of round ductwork is less than rectangular. Figure 19 shows the relative weight of rectangular duct to round duct for duct pressures from ±125 to ±2500 Pa when the equivalent diameter of the rectangular duct is the same as the round duct diameter. Equivalent diameter is defined as the diameter of a rectangular duct that has equal resistance to flow for equal flow and length.
Perimeter of round ducts is less than rectangular ducts. For rectangular duct aspect ratios from 2 to 4, the increase is approximately 30 to 55%. This increase results in increased insulation, including possible thickness to offset the additional heat transfer. For a comprehensive study of round and rectangular ducts as they affect system performance, consult McGill (1988).
Round ducts have an excellent resistance to low-frequency break-out noise (Schaffer 2011).
Duct rumble can occur in rectangular duct systems (Paulauskis 2016).
Space Constraints. Space constraints and obstructions, particularly ceiling height, are frequent problems. In these cases, the choice is round, rectangular, or flat oval ducts, depending on the air quantity that needs to be conveyed by the duct.
and Figure 20 cover three design cases: 0.65, 2, and 5 Pa/m friction rates. Friction rate 2 Pa/m is at the middle range. In Table 9, six ceiling (plenum) heights ranging from 400 to 1100 mm, in approximately 200 mm increments, are covered. Space allocated for insulation and reinforcement is 50 mm all around. The aspect ratio of the rectangular and flat oval ducts is 2:1. Figure 20 shows the maximum airflow as a function of the design friction rates and plenum spaces, as noted. For example, the 900 mm plenum as a design friction rate of 2 Pa/m is 6360 L/s maximum for a 800 mm round duct; 14 500 L/s for a 1600 × 800 mm rectangular duct; and 16 300 L/s for a 1600 × 800 mm flat oval duct. The airflow capacity of two round 800 mm ducts is 6360 L/s each.
When selecting a rectangular or flat oval duct, consider the following:
Rectangular duct has the advantage in mass because construction standards for flat oval exist only for +2500 Pa, whereas rectangular has seven pressure classes, starting at 125 Pa. All rectangular pressure classes are ±.
Note: Negative-pressure flat oval duct systems can be designed by using +2500 Pa sheet gages with the negative-pressure rectangular reinforcement welded to the duct.
Low-frequency breakout noise for flat oval is good, fair for rectangular (Schaffer 2011).
Duct rumble can occur in rectangular duct systems (Paulauskis 2016).
Perimeter. Using rectangular duct instead of flat oval when sized for equivalent diameter increases the perimeter roughly 17 to 7% for aspect ratios ranging from 1 to 4, respectively. This increase results in an increased surface area. Flat oval requires less insulation. For a comprehensive study of flat oval and rectangular ducts as they affect system performance, consult McGill (1995).
Duct Lengths. Rectangular duct is available in 1200, 1500, or 1800 mm lengths. Spiral round and flat oval can be provided in longer lengths.
Each air duct system should be tested, adjusted, and balanced. Guidance and procedures are given in Chapter 39 of the 2019 ASHRAE Handbook—HVAC Applications and in ASHRAE Standard 111. To determine fan total (or static) pressure from field measurements taking into account fan system effect, consult AMCA (2011b) Publication 203, which provides numerous fan/system configurations encountered in the field.
Many VAV noise complaints have been traced to control problems. Although most problems are associated with improper installation, many are caused by poor design. The designer should specify high-quality fans or air handlers within their optimum ranges, not at the edge of their operation ranges, where low system tolerances can lead to inaccurate fan flow capacity control. Also, in-duct static pressure sensors should be placed in duct sections having the lowest possible air turbulence (i.e., at least three equivalent duct diameters from any elbow, takeoff, transition, offset, or damper).
6.2 DESIGN RECOMMENDATIONS
Engage the architect and structural engineer early to coordinate shafts for systems.
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Commentary. Consult Noise and Vibration Control for HVAC Systems (Schaffer 2011) for the following guidelines related to duct systems: Selection of mechanical room walls Noise control for mechanical rooms Upward noise control for mechanical rooms Duct penetrations through walls Structural support of rooftop equipment for vibration control
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Route ducts as straight as possible to reduce pressure loss, noise, and first costs.
Use round spiral ducts whenever round ducts can fit within space constraints.
Avoid consecutive fittings and close-coupled fittings because they can significantly increase pressure losses.
Use return air plenums when possible because they reduce both energy and first costs. Plenum return requires fire-rated construction.
Design air distribution systems to minimize flow resistance and turbulence. High flow resistance increases fan pressure, which results in higher noise being generated by the fan, especially at low frequencies. Turbulence also increases flow noise generated by duct fittings and dampers, especially at low frequencies.
Efficient fittings create the least turbulence and noise. Figure 21 provides generalized guidelines for minimizing regenerated noise from takeoffs. Tables 10 and 11 provide specific guidance for tees and wyes.
Duct transitions should not exceed an included angle of 15°.
To avoid fan system effects, fans should discharge into duct sections that remain straight for as long as possible, up to 10 duct diameters from the fan discharge to allow flow to fully develop (Figure 22). Use the ASHRAE Duct Fitting Database (ASHRAE 2016) to account for fan outlet system effects (SD7 and SR7-series).
Design duct connections at the fan inlet for uniform and straight airflow. Both turbulence and flow separation at the fan blades can significantly increase fan-generated noise. To account for fan inlet system effects, use the ASHRAE Duct Fitting Database (ASHRAE 2016) (ED7 and ER7 series).
For all except very-noise-sensitive applications, select VAV reheat boxes for a total pressure loss from 125 to 150 Pa; for a fan-powered VAV box, from 150 to 175 Pa. For details, see Taylor and Stein (2004).
VAV terminal unit inlet duct should be the same size as the inlet to the box, unless the box is in the critical path or the length exceeds about 4.5 m from the takeoff. Duct upstream of box inlets should be rigid sheet metal duct, 1.2 m minimum. Do not use flexible duct immediately upstream of VAV boxes.
See Figure 23 for Schaffer’s (2011) guidelines for the installation of single-duct, dual-duct, and induction terminal units, as well as parallel and series flow fan-powered units.
Place fan-powered mixing boxes away from noise-sensitive areas.
Use demand-based static pressure set-point reset to reduce fan energy and noise.
For constant-volume systems, select the fan to operate as near as possible to its rated peak efficiency. Also, select a fan that generates the lowest possible noise at required design conditions. Using an oversized or undersized fan that does not operate at or near rated peak efficiency can substantially increase noise levels. For VAV applications, see Schaffer (2011).
Equal Friction Method. The equal friction method for sizing duct systems uses a constant friction rate. The target velocity determines the size of the first duct section both downstream and upstream of the fan. From the size determined by the target velocity, the design friction rate is determined to size all remaining duct sections except for connections to VAV and constant-volume (CV) terminal units and diffusers in the critical path, or for sections whose length exceeds around 4.5 m. In these cases, the inlet to terminal units should have at least three diameters of rigid duct and 2 m maximum of rigid or flexible duct (see Figure 8) to diffusers. The section upstream of the rigid duct to terminal units and the section of rigid/flexible duct to diffusers should be sized by the design friction rate, and an appropriate transition placed between sections. Refer to sections 8 and 9 in Figure 24 for an example.
Static Regain Method. The static regain method uses the conservation of momentum principle, which results in an increase in static pressure when the velocity is reduced in an airstream. This method sizes the main and branch ducts after a junction so that the recovery in static pressure caused by the reduced velocity is approximately equal to the total pressure drop caused by the duct and fittings in the subject duct section according to Equation (41). This is done iteratively by selecting a size for a section and checking to see if the section’s static pressure loss is close to zero. The basic steps are to design the first section at the velocity recommended in Table 12. Then each downstream section is subsequently sized using the same duct size as the upstream section as a starting point. If there is no static regain, meaning there is a positive static pressure loss, then the downstream section will be the same size as the upstream section. But, if there is static regain, meaning the change in static pressure is negative, then an opportunity exists to use a smaller duct section. Typically, the next smaller size available that is smaller than the upstream section is selected and the calculations are repeated. Refer to column 12 of Table 13 for examples of static regain.
For connections to VAV and CV terminal units and diffusers in the critical path or the section length with the terminal unit or diffuser exceeds around 4.5 m, the inlet to terminal units should have at least three diameters of rigid duct and 2 m maximum of rigid or flexible duct (see Figure 8) to diffusers. The section upstream of the rigid duct to terminal units and the section of rigid/flexible duct to diffusers should be sized by the static regain method and an appropriate transition placed between sections. Refer to sections 8 and 9 in Figure 24 for an example.
Rearranging,
where
| pv1 = velocity pressure at section 3, Pa |
| pv2 = velocity pressure at section 2, Pa |
| Δ pt,1–2 = total pressure loss from section 3 to 2, Pa |
| Δ ps,1–2 = static pressure regain (+) or loss (–), Pa |
Balancing. When system unbalance is greater than about 30%, it is recommended that balance be improved by changing duct size, and/or fittings from a select few that have comparable efficiency without introducing additional significant turbulence (noise). Tables 10 and 11 are examples.
Understanding what noise is and how to control it is fundamental to duct design. The underlying principles of sound and vibration are covered in Chapter 8. Chapter 49 of the 2019 ASHRAE Handbook—HVAC Applications contains technical discussions and design examples helpful to the design engineer. AHRI Standard 885 has procedures for estimating sound pressure levels in the occupied zone for the portion of the system downstream of terminal units. For guidance in designing HVAC systems to avoid noise and vibration problems, consult Schaffer (2011). Specifying quiet equipment and designing systems to avoid noise and vibration problems are necessary parts of the design process.
Noise in system comes from fans and generated noise resulting from air turbulence in ducts, individual fittings, close-coupled fittings, dampers, air modulating equipment and diffusers/grilles. Objectionable self-generated duct noise can be controlled by limiting the duct velocity to the values listed in Table 12.
The goal of duct design is an air distribution system without objectionable noise and minimum life-cycle cost (LCC). Noise can be controlled by limiting the duct velocity to the values listed in Table 12. For example, for ductwork above a suspended acoustical ceiling and with a maximum allowable NC or RC of 35 in the adjoining space, the maximum round and rectangular duct airflow velocities are 15.2 and 8.9 m/s, respectively. Use this velocity to size the first duct section either upstream or downstream from the fan for all duct design methods. Lowering duct velocity reduces system operating cost and minimizes noise from sources other than ducts (fittings, close-coupled fittings, and dampers).
|
Commentary: In 1988, Dr. Robert J. Tsal developed the T-method (Tsal and Adler 1987; Tsal et al. 1988, 1990) to design systems with a minimum LCC. The T-method was technically sound, but the system cost was not sufficiently accurate. As a result, the T-method was removed from the Handbook in 2013. |
Supply Duct Sizing.
Size fan supply ducts by either the equal friction (EF) or static regain (SR) method. The duct velocity anywhere in the system should not exceed the velocity listed in Table 12.
Size ducts downstream of terminal boxes, toilet exhaust ducts, and other low-pressure systems (e.g., in Figure 25C) using the equal friction method with a friction rate in the range from 160 Pa per metre such that the duct velocity in the duct anywhere in the system does not exceed the values in Table 12. Use Table 14 as a guide to select the design friction rate.
Terminal unit (VAV box) runouts should be full size with the exception that runouts in the “critical” path or a runout length greater than about 4.5 m should be a minimum 1.5 m of rigid duct full size and the remainder’s size determined by the design method (e.g., Example 8 and Figure 24).
Diffuser runouts should be full size, except for runouts in the critical path or a runout length greater than about 4.5 m: these runouts should be at least 2 m of flexible-duct full size, and the remainder’s size determined by the design method.
Return Duct Systems.
Sizing ducted returns depend on economizer relief system. For systems with return fans (Figure 25A), the return air ducts are sized using the EF method. The SR method does not work for negative-pressure duct systems.
If building relief is by relief fans or gravity/motorized dampers, pressure drop should be kept low as shown by the total pressure grade line associated with Figures 25B and 25C.
Control dampers in the economizer system should be selected (parallel or opposed blade) and sized in compliance with ASHRAE Guideline 16. Louvers should be sized in accordance with the section on Louvers, under Duct Design.
Unducted return air shafts are typically sized for low pressure loss using either a fixed friction rate, velocity, or both. Typically, shafts are simply sized based on velocity. Maximum velocities are generally in the 4.1 to 6.1 m/s range through the free area at the top of the shaft (highest airflow rate).
Diffuser/grille runouts should be designed by the design method of choice and a transition located at the outlet. For outlets in the critical path, the outlet duct size should be increased. Registers should not be used because of the damper.
Example 8.
For the VAV system shown in Figure 25, design the duct system by both the equal friction (EF) and static regain (SR) methods, and compare the section duct sizes, total pressure required for each path, and the unbalance between paths.
The system is located in Denver (1655 m elevation) and the duct is spiral round, galvanized steel (absolute roughness ε = 0.12 mm). The duct system is located above a suspended acoustical ceiling, and the allowable background sound in the occupied spaces is NC-35. Terminals T1, T2, T3, and T4 (VAV boxes with a one-row hot-water coil) are 380 L/s. VAV box loss coefficients are 1.67.
Solution: Use the ASHRAE Duct Fitting Database (ASHRAE 2016) to determine the air density, which is used to identify the velocity pressure and fitting loss coefficients for fittings. Density is 0.978 kg/m3 (Figure 26). The maximum design duct velocity from Table 12 is 15.2 m/s. The EF and SR calculations are shown by Tables 13 and 15.
For both the EF and SR designs, the duct size for Section 2 is 375 mm determined using CD11-3 (Figure 27). Sizing the remaining section depends on the design method. For the equal friction design method, the system equal friction (also from CD11-3) is 4.2 Pa/m. Knowing the system design friction rate, the duct sizes are determined by CD11-4. Figure 28 is an example for Section 4. For the static regain method, all sections other than the terminal unit runouts are sized by iterating using Equation (41) (column 12 in Table 13). For example, the results of the iteration for Section 4 are summarized by Table 16. The solution is the diameter that gives the result from Equation (41) that is closest to zero: in this case, 350 mm.
Tables 17 and 18 summarize the results. The SR design is slightly better balanced, 19% compared to 23%, and the SR design requires 9% less energy. In addition, the SR sizes are slightly larger, and thus potentially less noisy. Both methods have static regain, but the SR design uses the regain more efficiently.
6.4 INDUSTRIAL EXHAUST SYSTEMS
Chapter 33 of the 2019 ASHRAE Handbook—HVAC Applications discusses design criteria, including hood design, for industrial exhaust systems. Exhaust systems conveying vapors, gases, and smoke are designed by the equal-friction method. Systems conveying particulates are designed by the constant velocity method at duct velocities adequate to convey particles to the system air cleaner. For contaminant transport velocities, see Table 2 in Chapter 33 of the 2019 ASHRAE Handbook—HVAC Applications.
Two pressure-balancing methods can be considered when designing industrial exhaust systems. One method uses balancing devices (e.g., dampers, blast gates) to obtain design airflow through each hood. The other approach balances systems by adding resistance to ductwork sections (i.e., changing duct size, selecting different fittings, increasing airflow). This self-balancing method is preferred, especially for systems conveying abrasive materials. Where potentially explosive or radioactive materials are conveyed, the prebalanced system is mandatory because contaminants could accumulate at the balancing devices. To balance systems by increasing airflow, use Equation (42), which assumes that all ductwork has the same diameter and that fitting loss coefficients, including main and branch tee coefficients, are constant.
where
| Qc = airflow rate required to increase Pl to Ph, L/s |
| Qd = total airflow rate through low-resistance duct run, L /s |
| Ph = absolute value of pressure loss in high-resistance ductwork section(s), Pa |
| Pl = absolute value of pressure loss in low-resistance ductwork section(s), Pa |
For systems conveying particulates, use elbows with a large centerline radius-to-diameter ratio (r /D), greater than 1.5 whenever possible. If r /D is 1.5 or less, abrasion in dust-handling systems can reduce the life of elbows. Elbows are often made of seven or more gores, especially in large diameters. For converging flow fittings, a 30° entry angle is recommended to minimize energy losses and abrasion in dust-handling systems. For the entry loss coefficients of hoods and equipment for specific operations, see Chapter 33 of the 2019 ASHRAE Handbook—HVAC Applications and ACGIH (2016).
|
Commentary: Fitting CD11-5 of the ASHRAE Duct Fitting Database (ASHRAE 2016) is available for sizing minimum-velocity ducts conveying particulates. |
Example 9.
For the metalworking exhaust system in Figures 29 and 30, size the ductwork and calculate fan static pressure requirement for an industrial exhaust designed to convey granular materials. Pressure-balance the system by changing duct sizes and adjusting airflow rates. Minimum particulate transport velocity for the chipping and grinding table ducts (sections 1 and 5, Figure 30) is 20 m/s. For ducts associated with the grinder wheels (sections 2, 3, 4, and 5), minimum duct velocity is 23 m/s. Ductwork is galvanized steel, with absolute roughness of 0.09 mm. Use sizes for which spiral duct is available (80, 100, 125, 140, 150, 160, 180, 200, 224, 250, 280, 300, 315, 355, 400, 450, 500, 560, 600, 630, 710, 800, 900, 1000, 1120, 1250, 1300, 1400, 1500, 1600, 1800, 2000, 2100, 2200, 2300, 2400, and 2500 mm).
The building is one story, and the design wind velocity is 9 m /s. For the stack, use design J shown in Figure 2 in Chapter 46 of the 2019 ASHRAE Handbook—HVAC Applications for complete rain protection; stack height, determined by calculations from Chapter 46, is 4.9 m above the roof. This height is based on minimized stack downwash; therefore, the stack discharge velocity must exceed 1.5 times the design wind velocity.
Solution: The following table summarizes initial duct sizes and transport velocities for contaminated ducts upstream of the collector. The 22.8 m /s velocity in sections 2 and 3 is acceptable because the transport velocity is not significantly lower than 23 m /s. For the next available duct size (160 mm diameter), the duct velocity is 28.8 m /s, significantly higher than 23 m/s.
Design calculations up through the junction after sections 1 and 4 are summarized as follows:
For (initial) design 1, the imbalance between section 3 and section 2 (or 3) is 383 Pa, with section 3 requiring additional resistance. Decreasing section 3 duct diameter by manufactured spiral duct sizes results in the least imbalance, 88 Pa, when the duct diameter is 200 mm (design 2). Because section 3 requires additional resistance, estimate the new airflow rate using Equation (42):
At 900 L /s flow in section 3, 130 Pa imbalance remains at the junction of sections 1 and 4. By trial-and-error solution, balance is attained when the flow in section 3 is 860 L /s. The duct between the collector and fan inlet is 355 mm round to match the fan inlet (340 mm diameter). To minimize downwash, the stack discharge velocity must exceed 13.5 m /s, 1.5 times the design wind velocity (9 m/s) as stated in the problem definition. Therefore, the stack is 355 mm round, and the stack discharge velocity is 14.5 m /s.
Table 19 summarizes the system losses by sections. The straight duct friction factor and pressure loss were calculated by Equations (18) and (19). Table 20 lists fitting loss coefficients and input parameters necessary to determine the loss coefficients. The fitting loss coefficients are from the ASHRAE Duct Fitting Database. Figure 31 shows a pressure grade line of the system. Fan total pressure, calculated by Equation (15), is 1992 Pa. To calculate the fan static pressure, use Equation (17):
where 192 Pa is the fan outlet velocity pressure. The fan airflow rate is 1440 L /s, and its outlet area is 0.081 m3 (260 by 310 mm). Therefore, the fan outlet velocity is 17.9 m/s.
Hood suction for the chipping and grinding table hood is 560 Pa, calculated by Equation (5) from Chapter 33 of the 2019 ASHRAE Handbook—HVAC Applications [Ps,h = (1 + 0.25) (451) = 560 Pa, where 0.25 is hood entry loss coefficient Co, and 451 is duct velocity pressure Pv a few diameters downstream from the hood]. Similarly, hood suction for each grinder wheel is 170 Pa.
where 0.4 is the hood entry loss coefficient, and 336 Pa is the duct velocity pressure.
|
Duct Section |
Design Airflow, L/s |
Transport Velocity, m /s |
Duct Diameter, mm |
Duct Velocity, m /s |
|---|
|
1 |
850 |
20 |
224 |
21.6 |
|
2, 3 |
290 each |
23 |
125 |
23.6 |
|
4 |
580 |
23 |
180 |
22.8 |
|
5 |
1430 |
23 |
180 |
23.2 |
|
Design No. |
D1, mm |
Δ p1, Pa |
Δ p2+4, Pa |
Imbalance, Δ p1 − Δ p2+4 |
|---|
|
1 |
224 |
411 |
794 |
−383 |
|
2 |
200 |
762 |
850 |
−88 |
|
3 |
180 |
1320 |
712 |
+609 |
|
Q1 = 850 L /s
Q2 = 290 L /s; D2 = 125 mm dia. |
Q3 = 290 L /s; D3 = 125 mm dia.
Q4 = 850 L /s; D4 = 180 mm dia. |
ASHRAE members can access ASHRAE Journal articles and ASHRAE research project final reports at technologyportal.ashrae.org. Articles and reports are also available for purchase by nonmembers in the online ASHRAE Bookstore at www.ashrae.org/bookstore.
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