ASHRAE HVAC Noise Control: Handbook Chapter Summary for Acoustic Designers

HVAC noise is the dominant source of background noise in most commercial, institutional, and high-performance residential buildings. Unlike external environmental noise (which is attenuated by the building envelope) or occupant noise (which is attenuated by partitions), HVAC noise is generated within the building and delivered directly to occupied spaces through ductwork. Getting mechanical noise control right requires combining acoustical engineering principles with HVAC system design — a discipline that sits at the interface between architecture, mechanical engineering, and acoustics.

The primary reference for HVAC noise control in North American practice is Chapter 49 (Sound and Vibration Control) of the ASHRAE Handbook — HVAC Applications. Published by the American Society of Heating, Refrigerating and Air-Conditioning Engineers, the handbook chapter is updated with each edition of the Applications volume (typically every 4 years; the current relevant edition is 2019 ASHRAE Handbook Applications). This guide summarizes the key concepts, calculation procedures, and design criteria from that chapter in a format accessible to acoustic designers who may not routinely work with HVAC system design calculations.

Noise Criteria Curves: NC, RC, and NCB

NC Curves (Leo Beranek, 1957)

The NC (Noise Criteria) curves are the oldest and most widely referenced background noise rating system. Each NC curve is a set of octave-band maximum sound pressure levels (in the range 63 Hz to 8,000 Hz) that define a background noise spectrum considered acceptable for a given space type. The NC number (NC-20, NC-30, NC-40, etc.) represents the tangency point of the measured or predicted spectrum with the family of NC curves.

NC curves have important limitations:

• They do not address sub-63 Hz noise, yet many HVAC systems produce significant energy below 63 Hz
• They treat all spectra that fall below a given NC curve as equivalent, even if one is rumble-dominated and another is hiss-dominated
• They do not distinguish between steady background noise and intermittent tonal noise

RC Curves (W.E. Blazier, 1981)

The RC (Room Criteria) curves were developed specifically to address the limitations of NC curves for HVAC noise assessment. The RC system:

• Uses the same family of octave-band curves but introduces a quality assessment that categorizes the noise character
• Defines three regions: neutral (N — acceptable), rumble (R — low-frequency excess), and hiss (H — high-frequency excess)
• Extends the assessment to 16 Hz and 31.5 Hz octave bands to capture low-frequency rumble from HVAC systems

RC Mark II and NCB

Subsequent revisions produced RC Mark II and NCB (Balanced Noise Criteria) as refinements addressing additional limitations of the original RC methodology. ASHRAE Chapter 49 presents all three systems and recommends RC Mark II as the preferred assessment method.

ASHRAE Recommended RC Design Criteria by Space Type

HVAC Noise Sources: A Taxonomy

Before calculating duct noise, understanding the sources is essential. ASHRAE Chapter 49 categorizes HVAC noise sources as:

1. Fan Noise

Fans are the primary source of HVAC system noise. Fan noise has two components:

• Aerodynamic noise: broadband noise from turbulence in the airflow, plus tonal noise at the blade pass frequency (BPF = fan speed in RPS × number of blades)
• Mechanical noise: bearing noise, drive motor noise, imbalance-induced vibration

2. Terminal Unit Noise

VAV (Variable Air Volume) terminal units contain a pressure-independent airflow control damper and sometimes a reheat coil and fan. They generate noise in two ways:

• Regenerated noise from the control damper: as the damper closes to reduce airflow, turbulence increases and noise increases. At low damper positions (highly throttled), terminal unit noise can increase by 15–20 dB compared to fully open conditions
• Fan noise (for fan-powered VAV terminals): the integral fan generates noise at its operating speed

3. Diffuser and Grille Noise (Air Distribution Terminal Noise)

Air outlets (supply diffusers, return air grilles) generate noise when air exits the device at velocity. The primary noise-generating mechanism is shear turbulence as the high-velocity airflow from the device merges with room air. ASHRAE provides NC curves at various airflow velocities for common diffuser types and sizes; manufacturers provide octave-band Lw data for their products.

The design principle is simple: for noise-sensitive spaces, size diffusers for low velocity. For NC-30 or RC-30 compliance, supply velocity at the neck of the diffuser should typically be ≤ 2.5 m/s, and free area velocity should be ≤ 0.3 m/s for ceiling diffusers. Meeting these velocity limits while providing adequate throw for temperature uniformity sometimes requires more and smaller diffusers rather than fewer larger ones.

4. Duct Self-Noise (Flow-Generated Noise)

Even without fans or terminal units, turbulent airflow in ductwork generates self-noise at elbows, transitions, branches, and obstructions. ASHRAE provides equations for duct self-noise in straight sections and at fittings. Self-noise is most significant at:

• High duct velocities (>6 m/s in main ducts, >4 m/s in branch ducts)
• Sharp elbows without turning vanes
• Abrupt area transitions
• Dampers and fire dampers partially open

The ASHRAE Duct Noise Calculation Procedure

Step-by-Step Process

ASHRAE Chapter 49 presents a structured calculation worksheet for predicting the octave-band sound pressure level at a listener position from a duct system. The complete procedure:

Step 1: Fan sound power level Obtain octave-band Lw (63–8,000 Hz) from manufacturer at the design operating point. Add system effect corrections for outlet conditions (duct immediately at fan outlet reduces effective performance).

Step 2: Duct end reflection loss At low frequencies, sound energy reflects back toward the source at duct terminations rather than propagating into the room. End reflection loss (ERL) is largest at low frequencies:

Step 3: Duct wall transmission loss (breakout) Some sound energy in the duct breaks out through the duct walls. For rectangular sheet metal ducts, the breakout transmission loss depends on duct dimensions and frequency. Larger ducts have lower breakout TL. Breakout noise is often the dominant noise path for low-frequency fan noise in systems with large main ducts running through occupied spaces.

Step 4: In-duct attenuation Sound attenuates as it travels through ductwork. Unlined rectangular ducts provide approximately:

Lined ducts provide substantially more attenuation, particularly at mid and high frequencies.

Step 5: Fitting attenuation Elbows, junctions, and plenum boxes provide attenuation. A typical single-thickness elbow without turning vanes in a 300 mm duct provides approximately 2–5 dB attenuation at 500–2,000 Hz. Acoustically lined elbows with turning vanes can provide 8–15 dB attenuation.

Step 6: Terminal unit regenerated noise Add the octave-band Lw from the terminal unit at the design minimum airflow position (worst case noise).

Step 7: Diffuser noise Add the octave-band Lw from the diffuser at design supply airflow.

Step 8: Room correction Convert the combined source sound power at the room to a sound pressure level at the listener position. The ASHRAE formula uses the room constant R and the distance from the diffuser:

Lp = Lw + 10 log₁₀(Q/(4πr²) + 4/R)

Where Q is the directivity factor (2 for a diffuser in a ceiling — hemispherical radiation), r is the distance from diffuser to listener, and R is the room constant (m²).

Step 9: Compare with RC/NC criterion Plot the resulting octave-band spectrum against the target RC or NC curve. Identify which octave bands exceed the criterion and design remedial measures for those bands.

Silencer Design Principles

When the duct noise calculation indicates that the design criterion will be exceeded without additional treatment, silencers (also called attenuators or sound traps) are inserted into the ductwork.

Types of Silencers

Lined duct sections: the simplest form of silencer is a duct section with acoustic lining (typically 25–50 mm of mineral wool or fiberglass, density 24–48 kg/m³, covered with perforated metal facing). Lined duct provides approximately 1–3 dB/m attenuation at mid-high frequencies. It is most effective above 250 Hz and limited below 125 Hz.

Packaged rectangular silencers: proprietary rectangular attenuators with internal baffles of acoustic absorption material. Available in standard lengths (300–1,800 mm) and with various baffle thicknesses. Manufacturer data provides dynamic insertion loss (DIL) in octave bands for standard configurations.

Circular pod silencers: for circular ductwork, pod silencers use a central absorptive cylinder surrounded by an annular airway. Efficient in smaller diameters.

Reactive (tuned) silencers: expansion chambers and resonator silencers tuned to specific low-frequency problem frequencies. Unlike absorptive silencers, reactive silencers reflect rather than absorb sound energy, making them effective below 125 Hz where absorptive materials are inefficient.

Silencer Selection Criteria

When selecting a packaged silencer, three parameters must be balanced:

1. Required insertion loss at each problem octave band (from the noise calculation)
2. Pressure drop — silencers add resistance to airflow, increasing fan power and energy consumption. ASHRAE recommends that silencer static pressure drop not exceed 25–50 Pa for primary air silencers in energy-sensitive applications
3. Self-noise — silencers generate their own noise from airflow through the absorptive material. At high airflow velocities (>6–8 m/s face velocity), silencer self-noise can exceed the insertion loss benefit. The design point should keep face velocity below 5 m/s for near-source silencers in noise-sensitive spaces

Vibration Isolation

Structure-borne noise from HVAC equipment (fans, pumps, chillers, cooling towers) is transmitted through the building structure and re-radiated as airborne sound in occupied spaces. ASHRAE Chapter 49 covers vibration isolation design for rotating equipment.

Isolation Efficiency

The vibration transmissibility of a spring or elastomeric isolator depends on the ratio of the forcing frequency to the isolator's natural frequency:

η = 1 − 1/(1 − (fn/f)²)

Where η is the isolation efficiency (0–1), fn is the isolator natural frequency, and f is the forcing frequency. For effective isolation, the ratio f/fn should be at least √3 (approximately 1.73), which gives 50% isolation. For high-performance applications, f/fn > 5 is recommended for >95% isolation.

For a 1,500 RPM fan (forcing frequency 25 Hz), an isolator natural frequency of 5 Hz (soft spring, typically 25–50 mm static deflection) gives f/fn = 5, appropriate for vibration isolation in noise-sensitive buildings.

Common HVAC Vibration Isolation Details

Low-Frequency HVAC Noise: The Difficult Problem

Low-frequency noise (63 Hz and below) from HVAC systems is the hardest problem in mechanical acoustic design. At 63 Hz, the wavelength is approximately 5.5 m — comparable to room dimensions. This means:

• Absorptive silencers are essentially ineffective (absorptive materials must be at least λ/4 thick for meaningful absorption, which is 1.4 m at 63 Hz — impractical)
• Room absorption is nearly zero at 63 Hz in typical rooms
• The calculated NC/RC number underestimates the subjective annoyance because NC and RC curves are less sensitive to low-frequency rumble than the human hearing system
• Structural flanking paths become dominant at low frequencies, bypassing duct-borne transmission paths

• Select fans with low sound power at 63 Hz (axial fans are often worse at low frequency than centrifugal fans)
• Increase fan speed while reducing airflow volume to shift blade pass frequency above 63 Hz
• Use active noise cancellation in main ducts (specialized, expensive, limited effectiveness)
• Reactive expansion chambers tuned to 63 Hz (requires large duct volume)
• Vibration isolation of the complete AHU from the building structure to prevent structure-borne radiation

Integration with AcousPlan

AcousPlan's Room Acoustic Simulator includes NC and RC background noise assessment alongside reverberation time calculations. Enter the estimated HVAC background noise level for the space type, and the compliance view displays the NC and RC curves with pass/fail status against the ASHRAE recommended criteria for the selected building type. Space types from the ASHRAE table above are mapped to the building type selector in the simulator.

For mechanical engineers, the NC/RC calculator accepts octave-band SPL data directly and returns the NC rating, RC rating, and quality designation (N/R/H), with recommendations for the most cost-effective remediation path when the criterion is exceeded.

All calculations are advisory and must be verified by a qualified mechanical acoustic engineer against project-specific HVAC system characteristics.