Aluminum Case Acoustic Noise Control Engineering: Sound

A medical imaging equipment manufacturer shipped 12 portable ultrasound systems (each in a standard aluminum transit case, 600×400×300mm, 2.0mm 5052-H32 aluminum, with a standard non-acoustic EPDM gasket, 4 compression latches, and no acoustic treatment inside or outside) to a hospital in Singapore, for use in the hospital’s emergency department (ED) and intensive care unit (ICU). The ultrasound systems were high-end portable units ($45,000 each), with a built-in 15-inch LCD display, a solid-state drive (SSD), a lithium-ion battery (for portable use), and a high-voltage power supply (for the ultrasound transducer). The cases were designed as standard transit cases (for transport and storage), not as acoustic enclosures—so no acoustic noise control was considered during the case design. When the ultrasound systems were deployed in the ICU (a quiet patient care environment, with an ambient noise level of 35-40 dBA per hospital design guidelines), the hospital staff immediately complained: (1) the ultrasound system’s internal cooling fan (a 40mm × 40mm × 10mm DC brushless fan, running at 8,000 RPM, producing 38 dBA at 1m in free air—measured by the manufacturer in an anechoic chamber) was clearly audible in the quiet ICU environment, and the aluminum case was amplifying the fan noise (the case’s thin aluminum panels were resonating at the fan’s blade-pass frequency (8,000 RPM × 7 blades = 933 Hz, and its harmonics at 1,866 Hz, 2,800 Hz, etc.), acting like a speaker cone and radiating the fan noise outward—measured at 48 dBA at 1m outside the case, which is 10 dB louder than the fan’s free-air rating, due to the case’s panel resonance and the case’s lack of acoustic damping), (2) the fan noise was disturbing patients (the ICU has strict noise guidelines—WHO recommends <35 dBA at night in patient rooms, and the 48 dBA ultrasound fan noise was 13 dB above the guideline, causing patient complaints about difficulty sleeping and increased stress—3 patients in the ICU specifically complained about the noise during the first week of deployment), (3) the fan noise was interfering with clinical communication (the ED and ICU staff reported that they had to raise their voices to communicate over the fan noise, and that the noise was distracting during patient examinations—especially during cardiac ultrasound, where the sonographer needs to hear the patient's heart sounds and communicate with the patient), (4) the case was also transmitting structure-borne noise (the fan was mounted directly to the case's aluminum chassis with 4 metal screws, with no vibration isolation—so the fan's mechanical vibration (at 8,000 RPM = 133 Hz, and its harmonics) was transmitted directly through the screws to the aluminum case, causing the entire case to vibrate and radiate noise—this is structure-borne noise, which is even harder to control than airborne noise, because it travels through the solid structure with very little attenuation), (5) in one case, the fan noise was so loud that the hospital's biomedical engineering department measured it at 52 dBA at 1m (in a specific corner of the ICU, where the case was placed against a wall, causing a corner-loading effect that amplified the low-frequency noise by 4 dB), and the hospital threatened to return all 12 systems if the noise problem wasn't resolved within 2 weeks. The root cause analysis (conducted by the manufacturer's acoustic engineer) found: (1) the aluminum case provided no acoustic isolation (a bare 2.0mm aluminum panel has a transmission loss (TL) of only 15-20 dB at 100-1,000 Hz (per the mass law: TL = 20 × log10(m × f) – 47, where m is the surface mass in kg/m², f is the frequency in Hz—for 2.0mm aluminum, m = 5.4 kg/m², so at 500 Hz, TL = 20 × log10(5.4 × 500) – 47 = 20 × log10(2700) – 47 = 20 × 3.43 – 47 = 68.6 – 47 = 21.6 dB—only 21.6 dB isolation at 500 Hz, which is inadequate for a quiet environment), and the case's seams (lid-to-body, cable penetrations) provided additional noise leaks, reducing the effective TL to 10-15 dB), (2) the case had no acoustic damping (the bare aluminum panels were free to resonate at their natural frequencies—the 600×400mm panel has a fundamental resonance at approximately 80-120 Hz (depending on boundary conditions), and higher-order modes at 200-2,000 Hz—these resonances amplified the fan noise by 5-15 dB at specific frequencies, and the lack of damping meant the resonances were high-Q (narrow-band, high-amplitude), (3) the fan had no vibration isolation (the fan was mounted directly to the aluminum chassis with metal screws, with no rubber grommets or vibration isolators—so 100% of the fan's mechanical vibration was transmitted to the case, causing structure-borne noise), (4) the case had no internal acoustic absorption (the case interior was bare aluminum, with no acoustic foam or fiberglass lining—so the fan noise bounced around inside the case (reverberation), increasing the internal sound pressure level by 3-6 dB, and the reverberant field drove the case panels from the inside, increasing the outward noise radiation), (5) the case had no acoustic seals (the standard EPDM gasket was designed for weather sealing (water/dust), not acoustic sealing—EPDM is a good weather seal but a poor acoustic seal, because it's relatively hard (60-70 Shore A) and doesn't conform well to surface irregularities, leaving micro-gaps that allow noise to leak through—an acoustic gasket should be softer (30-40 Shore A) and more compressible, to fill all micro-gaps and provide a continuous acoustic seal). The total cost of this noise problem: (1) 12 systems at risk of return ($45,000 × 12 = $540,000 in potential revenue loss), (2) emergency acoustic re-design and re-work (engineering time + materials + re-work of 12 cases = $18,000), (3) hospital relationship damage (the hospital was a major account, with 5 facilities and $2M/year in potential purchases—the noise problem threatened the entire account relationship), (4) brand reputation damage (the hospital's biomedical engineering department reported the noise issue to the hospital's group purchasing organization (GPO), which could have affected the manufacturer's GPO contract status). The re-designed acoustic case (with proper noise control) cost an additional $65 per case: (1) acoustic damping layer (2mm butyl rubber-based damping compound, applied to all exterior panels—$15 per case), (2) vibration isolators for the fan (4 rubber grommets, 40 Shore A, isolating the fan from the chassis—$3 per case), (3) internal acoustic absorption (10mm melamine acoustic foam, applied to all interior walls (except the fan intake/exhaust areas)—$22 per case), (4) acoustic gasket (softer EPDM, 35 Shore A, with a higher compression ratio (35-45%), providing a better acoustic seal—$8 per case), (5) fan speed reduction (the fan was re-programmed to run at 6,000 RPM instead of 8,000 RPM (with a thermostatic control that increases speed only when the internal temperature exceeds 40°C)—reducing the fan noise by 6 dB (per the 6 dB per doubling of speed rule, approximately—8,000 to 6,000 is a 25% reduction, which gives about 3-4 dB reduction, plus the thermostatic control means the fan runs at low speed 80% of the time, giving an additional 2-3 dB average reduction)—$0 (firmware change, no hardware cost), (6) fan intake/exhaust silencing (acoustic foam lining in the fan duct, reducing the fan's tonal noise by 3-5 dB—$7 per case). Total additional cost: $55 per case (the damping $15 + isolators $3 + foam $22 + gasket $8 + silencing $7 = $55), or $660 for 12 cases. Compared to the $540,000 in potential revenue loss and $18,000 in emergency re-work, the acoustic upgrade would have paid for itself 850 times over. This is why acoustic noise control engineering matters—it's not just "for recording studios and concert halls," it's a critical consideration for any case containing noise-generating equipment (fans, pumps, compressors, motors, transformers, high-voltage power supplies) that will be used in noise-sensitive environments (hospitals, offices, recording studios, broadcast facilities, theaters, conference rooms, residential areas). Proper acoustic noise control engineering involves: acoustic theory (sound pressure level, frequency, mass law, transmission loss, damping, isolation, absorption), noise source identification (airborne vs. structure-borne, tonal vs. broadband), noise control strategies (source reduction, path control, receiver protection), acoustic materials (damping compounds, vibration isolators, acoustic foam/fiberglass, acoustic gaskets, silencers), and testing (sound pressure level per ISO 3744/3745, transmission loss per ASTM E90/ISO 10140, vibration measurement per ISO 10816). This guide covers all these aspects, with detailed formulas, calculation examples, and real-world case studies. Based on 15 years of designing acoustically-treated aluminum transit cases for medical imaging, broadcast/audio, industrial test & measurement, and office equipment applications.

Acoustic noise control engineering is one of the most overlooked aspects of aluminum transit case design—and one of the most common causes of customer complaints and product returns for equipment housed in transit cases. Most case buyers (and even some case manufacturers) treat acoustic noise control as a “recording studio only” requirement—”we don’t need acoustic treatment, our equipment is for industrial use.” But acoustic noise control is relevant for any case containing noise-generating equipment: (1) cooling fans (the #1 noise source in electronic equipment—40mm to 120mm DC/AC fans, running at 3,000-15,000 RPM, producing 30-60 dBA at 1m), (2) pumps and compressors (medical pumps, vacuum pumps, air compressors—producing 40-70 dBA, with strong low-frequency components (50-200 Hz) that are hard to isolate), (3) motors and actuators (servo motors, stepper motors, linear actuators—producing 35-55 dBA, with tonal components at the motor’s rotational frequency and its harmonics), (4) transformers and power supplies (high-voltage transformers, switching power supplies—producing 30-50 dBA, with 50/60 Hz hum and 20-100 kHz switching noise (the latter is often above human hearing but can cause audible effects through intermodulation)), (5) mechanical relays and contactors (producing impulsive “click” noise, 40-60 dBA peak, which can be annoying in quiet environments), (6) hard disk drives (HDDs—producing 30-40 dBA, with seek noise (2-5 kHz) and spindle noise (100-200 Hz)—though SSDs are eliminating this noise source in many applications), (7) fluid flow (air flow through vents/ducts, liquid flow through pumps/pipes—producing broadband noise, 30-50 dBA, especially at high flow rates). The common thread: if the equipment contains any moving parts or high-voltage/high-current components, it will generate noise—and if the equipment will be used in a noise-sensitive environment (hospital, office, recording studio, broadcast facility, theater, conference room, residential area, library, classroom), then acoustic noise control should be considered. And even if the equipment is “industrial” and “doesn’t need acoustic treatment,” adding a basic level of acoustic treatment (damping layer + vibration isolators + acoustic foam) is relatively inexpensive ($30-80 per case) and provides a significant noise reduction (10-20 dBA)—it’s cheap insurance against potential noise complaints and product returns. Proper acoustic noise control engineering involves: understanding acoustic theory (sound pressure level, frequency spectrum, mass law, transmission loss, damping, vibration isolation, acoustic absorption), identifying the noise source (airborne vs. structure-borne, tonal vs. broadband, dominant frequency), applying noise control strategies (source reduction—reduce noise at the source; path control—block/isolate/absorb the noise path; receiver protection—protect the receiver from noise), selecting appropriate acoustic materials (damping compounds, vibration isolators, acoustic foam/fiberglass, acoustic gaskets, silencers/mufflers), and verifying noise reduction through testing (sound pressure level per ISO 3744/3745, transmission loss per ASTM E90/ISO 10140, vibration measurement per ISO 10816). This guide covers all these aspects, with detailed formulas, calculation examples, and real-world case studies.

Acoustic Theory: Sound Pressure Level, Mass Law, and Transmission Loss

Before designing an acoustic noise control system, you must first understand the basic theory of acoustics and noise control. Sound is a mechanical wave that travels through a medium (air, water, solid) as a series of pressure variations (compressions and rarefactions). The key acoustic parameters for noise control are: (1) sound pressure level (SPL, measured in decibels, dBA for A-weighted, which approximates human hearing), (2) frequency (measured in Hertz, Hz—human hearing range is 20 Hz to 20,000 Hz, with the most sensitive range being 1,000-4,000 Hz), (3) sound power level (SWL, measured in dB re 1 pW—the total acoustic power radiated by a source, independent of distance and environment), (4) transmission loss (TL, measured in dB—the ability of a partition (panel, wall, case) to block sound transmission from one side to the other, measured in a laboratory with specific test conditions), (5) noise reduction (NR, measured in dB—the actual noise reduction between two spaces, which includes the TL of the partition plus the effect of absorption in the receiving room—NR = TL + 10 × log10(A/S), where A is the total absorption in the receiving room, S is the area of the partition).

Mass Law (the fundamental law of sound insulation): The mass law states that the transmission loss (TL) of a homogeneous, limp, infinite partition (like a panel) increases by 6 dB per doubling of mass (surface density, kg/m²) and by 6 dB per doubling of frequency. The mass law is the starting point for all acoustic insulation calculations—and it explains why thin aluminum panels (1.5-2.5mm, 4.1-6.8 kg/m²) have relatively poor TL at low frequencies (100-500 Hz), and why adding mass (thicker panels, lead lining, mass-loaded vinyl) improves TL. But the mass law has important limitations: (1) it applies only to homogeneous, limp, infinite partitions—real panels have stiffness (which causes coincidence effects and panel resonance), finite size (which causes edge effects and modes), and discontinuities (seams, holes, penetrations—which cause noise leaks that dominate the TL at high frequencies), (2) it assumes the panel is limp (no stiffness)—real aluminum panels have significant stiffness, which causes the “coincidence effect” (at the coincidence frequency, the bending wave speed in the panel matches the sound wave speed in air, causing a dip in TL—for 2.0mm aluminum, the coincidence frequency is approximately 5-10 kHz, where the TL can be 10-15 dB lower than the mass law prediction), (3) it assumes no panel resonance—real panels have natural frequencies (fundamental and higher-order modes) where the panel resonates and radiates sound very efficiently, causing dips in TL of 5-20 dB at those frequencies, (4) it assumes no leaks—real cases have seams, holes, and penetrations that can reduce the effective TL to 10-20 dB (regardless of the panel’s mass law TL), because noise leaks through these openings (a 1% open area can reduce the TL from 40 dB to 20 dB, and a 10% open area can reduce it to 10 dB).

Panel resonance and damping: Panel resonance is the #1 noise problem in aluminum transit cases—and it’s almost entirely preventable with a damping layer. Every panel has natural frequencies (resonant frequencies) where it vibrates very efficiently when excited by an external force (like a fan’s vibration or an internal sound field). At these resonant frequencies, the panel radiates sound very efficiently (like a speaker cone), amplifying the noise by 5-15 dB compared to non-resonant frequencies. The fundamental resonant frequency of a simply-supported rectangular panel is: f₁₁ = (π/2) × sqrt(D/(m × a⁴) × (1 + (a/b)²)²), where D is the bending stiffness (D = E × t³ / (12 × (1-ν²))), m is the surface mass, a and b are the panel dimensions. For a 600×400mm × 2.0mm aluminum panel (simply supported), the fundamental resonance is approximately 80-120 Hz, with higher-order modes at 200-2,000 Hz (these higher-order modes are the ones that typically coincide with fan blade-pass frequencies (500-2,000 Hz) and cause the most noise amplification). Damping is the process of adding a material to the panel that converts vibrational energy into heat (through viscoelastic losses), reducing the panel’s resonant response (the amplitude of vibration at resonance) and thus reducing the noise radiation. The effectiveness of damping is measured by the loss factor (η, dimensionless)—a higher loss factor means more damping, less resonant vibration. A bare aluminum panel has a very low loss factor (η ≈ 0.0001-0.001), meaning it rings like a bell at resonance (high-Q resonance, high amplitude). Adding a damping layer (like butyl rubber-based damping compound, η ≈ 0.1-0.5, or asphalt-based damping pad, η ≈ 0.05-0.2) can increase the panel’s effective loss factor to η ≈ 0.05-0.2, reducing the resonant amplitude by 10-30 dB (a factor of 3-30× reduction in vibration amplitude, which translates to 10-30 dB reduction in noise radiation at resonance). Damping is the single most cost-effective noise control measure for aluminum cases—a $10-20 damping layer (2mm butyl rubber compound applied to all exterior panels) can reduce case noise by 5-15 dBA (especially at the resonant frequencies where the case is loudest), which is often enough to resolve a noise complaint.

Vibration isolation (structure-borne noise control): Vibration isolation is the process of decoupling a noise/vibration source (fan, pump, motor, compressor) from the case structure, so that the source’s mechanical vibration is not transmitted to the case (which would cause the case to vibrate and radiate noise—this is structure-borne noise, which is often more difficult to control than airborne noise, because it travels through solid structures with very little attenuation). The principle of vibration isolation is the mass-spring system: the source (mass, m) is mounted on a spring (isolator, stiffness k), and the system has a natural frequency f_n = (1/(2π)) × sqrt(k/m). For isolation to occur, the excitation frequency (f_exc, the frequency of the source’s vibration) must be greater than √2 × f_n (f_exc > 1.414 × f_n). At f_exc = √2 × f_n, the transmission is 1 (no isolation, no amplification). At f_exc √2 × f_n, the isolator provides attenuation (isolation), with the transmission decreasing as f_exc/f_n increases (the higher the ratio, the better the isolation—at f_exc/f_n = 3, transmission is ~0.125 (18 dB isolation); at f_exc/f_n = 5, transmission is ~0.04 (28 dB isolation)). For a typical cooling fan running at 8,000 RPM (133 Hz rotational frequency, 933 Hz blade-pass frequency), a vibration isolator with f_n = 20-30 Hz would provide good isolation at both frequencies (f_exc/f_n = 4.4-6.7 at 133 Hz, and 31-47 at 933 Hz—excellent isolation at the blade-pass frequency, which is usually the dominant noise source). Common vibration isolators for transit case applications: (1) rubber grommets (the simplest and cheapest—4 rubber grommets (40-60 Shore A) between the fan and the chassis, isolating the fan’s vibration—f_n typically 20-40 Hz, providing 10-20 dB isolation at fan frequencies, cost $1-5 per fan), (2) rubber mounts (cylindrical or conical rubber mounts, for heavier sources like pumps and compressors—f_n typically 10-30 Hz, providing 15-25 dB isolation, cost $5-20 per mount), (3) spring mounts (coil springs, for very heavy sources or for very low-frequency isolation (f_n = 3-10 Hz)—providing 20-30 dB isolation, but springs have very low damping (η ≈ 0.001), which can cause resonant ringing at f_n—so spring mounts are usually used with a damping element (rubber pad in series with the spring), cost $10-50 per mount), (4) foam pads (open-cell polyurethane or melamine foam, for light sources and for high-frequency isolation—f_n typically 30-60 Hz, providing 5-15 dB isolation at high frequencies, cost $1-5 per pad). Our standard: For cooling fans (the most common noise source in electronic equipment), we use 4 rubber grommets (50 Shore A natural rubber or EPDM, with a metal washer for load distribution) between the fan and the chassis—this provides 10-20 dB vibration isolation at fan frequencies, at a cost of $1-3 per fan. For heavier sources (pumps, compressors, motors >1kg), we use cylindrical rubber mounts (40 Shore A, with threaded studs for mounting) with f_n = 10-20 Hz, providing 15-25 dB isolation. For critical applications (medical, recording, broadcast), we use spring-and-damper mounts (for very low f_n and high isolation), and we verify the isolation with vibration measurement (accelerometer on the case, before and after isolation, per ISO 10816).

Aluminum Cases for Engineered Equipment

After 15 years of designing acoustically-treated aluminum transit cases, we’ve learned that acoustic noise control is not a “recording studio only” requirement—it’s a critical consideration for any case containing noise-generating equipment (fans, pumps, compressors, motors, transformers, high-voltage power supplies) that will be used in noise-sensitive environments (hospitals, offices, recording studios, broadcast facilities, theaters, conference rooms, residential areas). A non-acoustic case (bare aluminum, no damping, no vibration isolation, no acoustic foam, standard weather gasket) provides 0-10 dBA noise reduction—meaning that a 40 dBA fan (measured in free air) can result in 45-55 dBA outside the case (due to panel resonance amplifying the noise by 5-15 dB)—which is too loud for any noise-sensitive environment. A properly acoustically-treated case (damping layer + vibration isolation + acoustic foam + acoustic gasket + fan speed control) provides 15-30 dBA noise reduction—reducing the same 40 dBA fan to 25-35 dBA outside the case—well below the noise threshold for most sensitive environments. And the cost is modest: a basic acoustic upgrade (damping + isolation + foam + acoustic gasket) adds $30-80 per case (5-10% of the case cost), while a medical-grade acoustic system (double-wall + double isolation + melamine foam + silencer + fan speed control + accredited testing) adds $200-500 per case (20-40% of the case cost). Compared to the cost of a noise complaint ($8,200 average per incident, including product returns, re-work, engineering, and customer relationship damage), acoustic noise control is cheap insurance. Proper acoustic noise control engineering involves: acoustic theory (SPL, frequency, mass law, TL, damping, isolation, absorption), noise source identification (airborne vs. structure-borne, tonal vs. broadband, dominant frequency), noise control strategies (source reduction, path control, receiver protection), acoustic materials (damping compounds, vibration isolators, acoustic foam/fiberglass, acoustic gaskets, silencers), and testing (SPL per ISO 3744/3745, TL per ASTM E90/ISO 10140, vibration per ISO 10816). We don’t just “add some foam and call it acoustic”—we engineer the complete acoustic noise control system for the specific application (equipment type, noise source, noise level, frequency spectrum, environment, noise requirement), and we verify performance through accredited testing.

We are an OEM/ODM partner working with partner factories in Changzhou’s aluminum case manufacturing cluster, operating under ISO 9001 quality systems. Our acoustic noise control engineering capabilities include:

• Acoustic consulting and specification: We help you determine the acoustic noise control requirements for your specific application (equipment type: fan-cooled electronics/pump/motor/compressor/transformer; noise source: airborne/structure-borne/both; noise level: measured or estimated; frequency spectrum: dominant frequencies; environment: hospital/office/industrial/outdoor; noise requirement: dBA at 1m, per ISO 3744 or customer spec), and we provide detailed acoustic specifications (damping material/thickness, isolator type/f_n, foam type/thickness, gasket type/compression, fan speed control, silencer design, test methods) with every acoustically-treated case design. We can also conduct acoustic assessments of your existing case fleet (identifying noise sources, measuring SPL, recommending upgrades).
• Acoustic damping layer design and application: We offer acoustic damping layers in all standard types: (1) butyl rubber-based damping compound (our standard—loss factor η = 0.2-0.4 at 20°C, 2mm thick, applied to exterior panels, cost $10-25 per case, providing 10-20 dB reduction at panel resonance), (2) asphalt-based damping pad (lower cost, η = 0.05-0.15, 1.5-3mm thick, self-adhesive, cost $5-15 per case, suitable for less demanding applications), (3) constrained-layer damping (CLD) system (two aluminum panels with a viscoelastic core in between—providing η = 0.3-0.6, the highest damping effectiveness, but also the highest cost and weight, used for critical applications like medical and recording, cost $30-80 per case), (4) spray-on damping (liquid-applied damping compound, sprayed onto panels, suitable for complex shapes and retrofits, cost $15-35 per case). We design the damping layer thickness and material based on the panel’s resonant frequencies and the required damping level, and we apply it with a roller (for sheet material) or spray gun (for liquid material) to ensure full contact and no air bubbles.
• Vibration isolation design and installation: We offer vibration isolators for all standard noise/vibration sources: (1) rubber grommets (for cooling fans and light sources <1kg—40-60 Shore A EPDM/natural rubber, f_n = 20-40 Hz, 10-20 dB isolation, cost $1-5 per source), (2) cylindrical rubber mounts (for pumps, compressors, motors 1-10kg—30-50 Shore A, f_n = 10-25 Hz, 15-25 dB isolation, cost $5-20 per mount), (3) conical rubber mounts (for heavier sources 10-50kg, with progressive stiffness—f_n = 8-20 Hz, 15-25 dB isolation, cost $10-30 per mount), (4) spring-and-damper mounts (for critical applications and very low-frequency isolation—coil spring + rubber damper, f_n = 3-10 Hz, 20-30 dB isolation, cost $20-60 per mount), (5) foam pads (for light sources and high-frequency isolation—open-cell polyurethane/melamine, f_n = 30-60 Hz, 5-15 dB isolation, cost $1-5 per pad). We design the isolation system based on the source's weight, excitation frequency, and required isolation level, and we verify the f_n with calculation (and with vibration measurement for critical applications). We also offer double-isolation (source on a sub-chassis, which is itself isolated from the main chassis) for critical applications, providing an additional 5-10 dB isolation.
• Internal acoustic absorption design and installation: We offer internal acoustic absorption in all standard types: (1) melamine foam (BASF Basotect, our standard for medical/broadcast/recording—density 9 kg/m³, NRC = 0.75, UL 94 V-0 fire rating, no VOC outgassing, 10-50mm thick, cost $15-40 per case), (2) polyurethane foam (standard for industrial/office—density 20-30 kg/m³, NRC = 0.6-0.8, UL 94 HF-1 fire rating, 10-50mm thick, cost $10-30 per case), (3) fiberglass (for high-temperature applications—density 32-48 kg/m³, NRC = 0.8-0.95, can withstand 200-500°C, must be encapsulated in a fabric bag (to prevent fiber irritation), 25-100mm thick, cost $20-50 per case), (4) acoustic fabric (for covering foam/fiberglass, providing a clean appearance and protecting the absorption material—acoustically transparent (must not impede sound transmission), fire-rated, cost $5-15 per case). We design the absorption type, thickness, and coverage based on the case’s internal volume, the noise source’s frequency spectrum, and the required reverberation reduction, and we apply it with adhesive (for foam) or mechanical fasteners (for fiberglass/heavy materials), keeping fan intake/exhaust and equipment mounting areas clear.
• Acoustic gaskets and sealing: We offer acoustic gaskets in all standard types: (1) soft EPDM (our standard—30-40 Shore A, softer than standard 60 Shore A for better conformability, 3-5mm cross-section, compressed 35-45%, providing a continuous acoustic seal with no micro-gaps, cost $5-15 per case), (2) silicone (for high-temperature applications—30-40 Shore A, -55°C to +200°C, cost $8-20 per case), (3) fluorosilicone (for fuel/oil-exposed applications—30-40 Shore A, fuel/oil resistant, cost $12-25 per case), (4) closed-cell foam (for very soft, high-compression sealing—polyethylene or polyurethane foam, 20-30 Shore A, compressed 50-60%, cost $3-10 per case), (5) acoustic caulk (for sealing small gaps and irregular openings—acrylic or silicone-based, paintable, cost $2-5 per case). We design the gasket type, cross-section, and compression based on the lid rigidity, latch configuration, and required sealing level, and we install it in a precision-machined groove (for consistent compression and alignment) or as a surface-mounted gasket (for retrofit applications).
• Fan silencers and airflow management: For fan-cooled equipment, we offer fan silencers and airflow management: (1) acoustic-lined fan duct (the most common—5-10mm melamine or polyurethane foam lining in the fan intake/exhaust duct, 50-100mm long, reducing tonal blade-pass noise by 4-8 dB, cost $5-15 per case), (2) reactive silencer (for low-frequency fan noise—expansion chamber with inlet/outlet pipes, using impedance mismatch to reflect low-frequency sound, reducing 60-250 Hz noise by 5-15 dB, cost $15-35 per case), (3) dissipative silencer (for high-frequency fan noise—acoustic foam lining in a long duct, absorbing high-frequency sound, reducing 500-10,000 Hz noise by 8-20 dB, cost $10-25 per case), (4) fan speed control (PWM thermostatic control, reducing fan speed when cooling demand is low—3-10 dBA average noise reduction, cost $0-5 (firmware/hardware), (5) fan selection assistance (we can recommend low-noise fans for your application—sleeve-bearing vs. ball-bearing vs. fluid-dynamic bearing, 3,000 vs. 5,000 vs. 8,000 RPM, 40mm vs. 60mm vs. 80mm—larger, slower fans are quieter for the same airflow, because noise increases with the 5th-6th power of fan speed (the “fan law”: L ≈ 50 × log10(N2/N1), so doubling speed increases noise by 15 dB). We design the silencer type, size, and foam thickness based on the fan’s airflow requirement, noise spectrum, and available space, and we verify the airflow resistance (pressure drop) to ensure adequate cooling.
• Acoustic testing and quality control: Every acoustically-treated case undergoes acoustic quality control inspection: (1) visual inspection (damping layer: full coverage, no air bubbles, no damage; foam: correct type/thickness, full coverage, no delamination; gasket: correct type, properly seated, no gaps; isolators: correct type, properly installed, no contact between source and chassis (no “short circuits” in the isolation path); silencer: properly installed, no airflow blockage), (2) gasket compression test (measuring compression at 5+ points along the seam, verifying 35-45% compression range), (3) isolation verification (visual check that the source is fully isolated from the chassis—no metal-to-metal contact, no cable short circuits, no rigid connections that bypass the isolators), (4) spot-check SPL test (for production runs, we test 1 in 10 cases with a portable sound level meter (Type 2, IEC 61672), at 1m, in a quiet room (<30 dBA background)—verifying that the SPL is below the specified requirement; for prototype and critical applications, we test 100% of cases), (5) vibration spot-check (for critical applications, we measure case panel vibration with a portable accelerometer (vibration meter, ISO 10816), verifying that the vibration velocity is below the specified threshold), (6) accredited acoustic testing (for prototypes and qualification units, we test at an accredited acoustic test lab: SPL per ISO 3744/3745 (hemi-anechoic chamber, 10 microphone positions), TL per ASTM E90/ISO 10140 (transmission loss suite, 1/3 octave bands from 50 Hz to 10 kHz), vibration per ISO 10816 (velocity and acceleration measurements)—we provide a full test report with SPL vs. frequency, TL vs. frequency, and vibration spectra). We maintain inspection records for every case for a minimum of 5 years, and we provide a Certificate of Acoustic Compliance with every shipment (documenting that the acoustic noise control system meets the specified requirements and passes all tests).

Typical lead times: 3-4 weeks for prototypes (with acoustic system design, material sourcing, and prototype SPL testing at an accredited lab), 5-8 weeks for production (25-250 cases). MOQ: 1 unit for custom acoustic design. Pricing: acoustic upgrade from “standard” (no damping, no isolation, no foam, standard weather gasket) to “basic acoustic” (damping layer + vibration isolators + acoustic foam + acoustic gasket) adds $30-80 per case (depending on case size and number of noise sources). “Medical-grade acoustic” (double-wall + double isolation + melamine foam + silencer + fan speed control + accredited SPL testing) adds $200-500 per case. Quality assurance: 2-year warranty against acoustic defects (damping delamination, isolator failure, foam degradation, gasket compression set—material and workmanship), with free re-work for any case that fails the specified SPL test under normal use (within the warranty period). For medical/recording/broadcast applications, we offer extended warranty (5 years) and enhanced testing (100% SPL test, 100% vibration test) for an additional charge.