Audio analysis,
explained plainly.

01 · What the analysis says

The current C1 design is a plausible, extensively screened prototype: its bass tuning is around 42 Hz, the added ties reduce panel motion in the modelled cases, and 25 mm acoustic lining reduces predicted internal ringing. The crossover matches the manufacturer’s published circuit.

It has not been acoustically measured. We cannot yet claim a flat full-range response, a particular maximum listening level, or that this acrylic cabinet sounds as good as the original Tifa 6. The original crossover’s acoustic behaviour depends on the cabinet, driver positions and real drivers.

02 · The design these results describe

This report consolidates the saved analyses through revision C1, dated 21 October 2026. C1 is the audited successor to revision C shown in the earlier project report. Earlier B studies are identified separately below.

The acrylic ties provide the main stiff load paths; the PETG fittings connect and locate them. The critical brace connections use through-bolts and locking nuts. C1 widened the cross-tie from 24 to 30 mm and adjusted its sockets to improve fastening margins.

Frosted acrylic and mandarin-orange PETG are appearance choices. No separate acoustic benefit was established for those finishes.

03 · Bass extension, port tuning and driver movement

A low-frequency electrical/mechanical/acoustic model combines the woofer’s published parameters with the cabinet air spring and port air mass. It assumes 1 m half-space radiation, coincident woofer/port output, linear behaviour and specified losses. It is not a complete room or full-range simulation.

Port lengths from 165 to 209 mm were screened. With the woofer network included, lengthening from 185 to 209 mm lowered tuning from 42.34 to 40.17 Hz, but moved the calculated −3 dB point from 51.90 to 52.70 Hz. It increased 30 Hz output by only about 0.76 dB at the screened voltage. The design retains 185 mm: lower port tuning alone is not the same as better bass extension.

The early simple dimensional tuning screen gave 40.75–44.05 Hz. The broader uncertainty study below gave a wider range because it also varied effective volume and other properties.

04 · Variation, protective filters and output limits

The broader bass study used 512 seeded parameter combinations, including driver variation, losses, port dimensions, coil resistance and component tolerance. Assumptions included effective volume from 95% to 120% of nominal, compliance ±15%, moving mass ±10%, motor strength ±5%, resistance ±10%, and woofer-coil resistance from 0.2 to 0.8 Ω. These were independent assumed ranges, not measured production statistics.

Sampled tuning ranged from 38.21 to 43.69 Hz. The table gives the lowest limits found in those samples over 20–200 Hz, with the woofer crossover included. Voltage is sinusoidal RMS at the amplifier output before the crossover.

35 Hz fourth-order Butterworth filtering and a provisional 7 V RMS test ceiling are a starting example from these models, not a certified operating limit or a music-power rating. The earlier nominal 8 V example omitted the passive network and sampled variation; it should not be interpreted as a universally safe limit. A 17 m/s air-speed screen is an illustrative comparison value, not a universal threshold for audible chuffing.

Random sampling is not exhaustive worst-case analysis. These results exclude clipping, nonlinear suspension, port turbulence, thermal damage and actual music crest factors.

05 · Timing, listening height and warming

No room simulation, real driver directivity, baffle-edge diffraction solution or measured acoustic phase was included. Those omissions matter especially around the crossover and in judging tonal balance.

06 · Crossover and amplifier load

The passive crossover was modelled electrically from 20 Hz to 20 kHz, with approximate driver impedances. A further 256 samples varied resistors, capacitors and inductors by ±5%, plus assumed coil resistance ranges. This checks electrical behaviour; it does not calculate the real combined acoustic response.

All 11 nominal component values match the manufacturer’s Tifa 6 schematic:

The manufacturer’s tweeter polarity is reversed relative to the woofer. An independent circuit solution matched the existing branch model across 301 frequencies with a maximum complex-voltage discrepancy of 1.04 × 10⁻¹⁵. That proves agreement of the two electrical implementations under their test loads, not that the real speaker will have a smooth crossover.

No crossover component changes are acoustically validated for this acrylic cabinet. Preserve the verified schematic as the prototype baseline, then use measured driver responses and impedance to decide whether tuning is needed.

07 · Component heating and internal electronics

The following are separate frequency-sweep dissipation maxima at 7 V RMS. They occur at different frequencies and must not be added into a single simultaneous total.

Actual purchasing SKUs, coil DC resistance, component dimensions, magnetic interaction and populated-board fit remain unverified. The 120 × 90 × 55 mm electronics allowance is a reserved space, not a qualified board layout. Keep lining clear of components that dissipate heat and confirm the physical arrangement before assembly. Coil orientation and separation need attention; they were not solved electromagnetically.

There is no internal amplifier, power supply or DAC in this passive design. Speaker return is the amplifier’s negative output, not an earth connection. Illustrative render wiring and component envelopes are not manufacturing evidence.

08 · Internal air resonances

The actual CAD air space was divided into small cells and solved for standing-wave patterns. Drivers and port were treated as closed rigid surfaces, so this is a cavity-mode study rather than a driven, open-port sound simulation.

Meshes of 14, 12 and 10 mm were compared. The 10 mm model retained 15,292 connected cells representing 15.292 litres, about 2.6% below the smooth CAD volume. Successive fine-grid changes were under 5% for the checked first six modes. A rectangular-box benchmark differed by at most 0.127%, and numerical residual checks passed.

The actual geometry’s lowest mode is approximately 433 Hz, versus about 476 Hz for the earlier empty rectangular height calculation. Real geometry matters. This solver does not include excitation strength, sound escaping through the drivers or port, flexible walls, or damping.

09 · Acoustic lining and internal ringing

A Miki porous-absorber model estimated the extra decay from five fitted lining patches. It used the CAD cavity pressure patterns, an assumed unlined resonance Q of 50, and nine combinations of thickness and airflow resistivity. Airflow resistivity describes how hard it is for air to move through the material; bulk density alone does not establish it.

For the lowest mode, around 433 Hz:

The flow-resistivity assumptions were 5,000, 10,000 and 20,000 Pa·s/m². At 25 mm and the middle value, the lowest mode’s Q is 19.76 and decay time is about 0.100 s. Its predicted pressure reduction is about 8.06 dB relative to the assumed unlined reference.

This is a first-order weak-loss approximation. Several more heavily damped higher modes exceeded its validity threshold and were deliberately withheld. The model does not reposition the acoustic boundary to the lining’s inner face, predict resulting frequency shifts, or solve coupled fabric/panel motion.

Five 25 mm lining pieces fit the current geometry, including notches around lower bracing, with checked hard-part clearance and at least 5 mm driver clearance. Keep the port clear. Porous lining is not treated as a solid block displacing its full bulk volume.

Absorption and structural damping are different. This lining absorbs internal sound. No measured reduction in acrylic wall vibration from constrained-layer damping has been established.

10 · Cabinet joints and side-to-side bracing

A plate finite-element model screened 15 cross-tie positions and three assumed joint stiffnesses. It included the tie’s axial stiffness and added mass, two pressure patterns, 18 vibration modes and a 100–1,500 Hz frequency sweep. Selected geometry was refined from an 8 mm to a 6 mm mesh.

The selected cross-tie is at Y = 110 mm, Z = 110 mm in the CAD coordinates. Results below compare the same model with and without this additional tie:

These percentages describe the largest calculated panel-motion peaks under chosen loads. They are not percentages of sound-quality improvement or cabinet-noise reduction. With soft joints, added mass can even lower the first resonance while slightly reducing the screened response peak.

The selected position performed well across the tested cases, but another location was almost tied. It is a practical choice within a limited search, not a proven global optimum. Joint stiffness has not been measured or guaranteed by the bolts.

The model assumes opposing walls move inward symmetrically, acrylic E = 3.2 GPa, and structural loss factor 0.03. It excludes complete cabinet coupling, driver reaction forces and modes that do not stretch the tie. Mesh checks passed for first-mode and response-ratio stability.

11 · Front-to-back ties

A separate coupled front/rear plate model included the large driver and port openings and lumped attached masses. Two acrylic ties sit at X = ±65 mm, Z = 70 mm. The study used 24 modes over 40–1,200 Hz, uniform symmetric pressure, assumed loss factor 0.03 and 8, 6 and 4 mm meshes.

The response-ratio change between the final mesh refinements was about 3.25%. These tie dimensions were unchanged from C to C1, so the existing study remains applicable to that geometry. Idealized edge supports, point masses and support springs limit its realism. It is not a whole-cabinet radiation prediction and omits detailed adapter stiffness and motor reaction forces.

12 · Earlier studies and what they contributed

These results explain the design progression. They must not be substituted for current C1 predictions.

The earlier plate models used ideal support boundaries, assumed driver/electronics masses and selected loads. Their analytical benchmarks, residual and mesh checks passed within the recorded thresholds. Those checks establish numerical consistency, not real-joint accuracy. The PETG cantilever stiffness estimate was only a comparison yardstick, not a qualification of the assembled socket.

13 · What was checked—and what remains unknown

The C1 package records 64 numerical and geometry checks. This is not 64 physical acoustic tests. Checks include circuit power/current balance, finite results, port formula agreement, filter response, mesh comparisons, absorber passivity, lining clearance and assembly geometry.

Other unmodelled effects include nonlinear distortion, turbulent port flow, thermal compression beyond resistance-only sensitivity, room interaction, real acrylic/PETG damping and creep, air leaks, component magnetic coupling and a coupled structural–acoustic simulation of the entire enclosure.

14 · Measurements to make on the first build

1. Confirm parts and fit. Measure the acrylic stock, test printed joints, select the real crossover components and check their board layout. The CNC release remains conditional on stock tolerances and physical fit.
2. Check wiring before connecting an amplifier. Verify the published circuit, driver polarity and absence of shorts. Record actual coil resistance.
3. Measure low-level input impedance. Check the two vented-box impedance peaks, tuning region and unexpected resonances. Investigate leaks or rattles before increasing level.
4. Measure woofer and port at close range. Combine with appropriate area and phase treatment to assess bass tuning; do not directly compare a raw near-field curve with a 1 m model.
5. Measure each driver and the combined speaker at a consistent reference position. Use time-windowed measurements where appropriate, plus vertical and horizontal angles. Establish acoustic phase and crossover integration before changing components.
6. Check cabinet motion and ringing. Use an accelerometer or suitable vibration sensor, repeat at matched excitation, and compare tied/untied or lined/unlined configurations only where safely reversible.
7. Increase level gradually while observing behaviour. Check distortion, port noise, cone motion, loose fittings and crossover temperature. Establish a measured protection setting and clean operating range.
8. Listen after the checks. Use familiar material and matched playback levels. Tune based on repeatable measurements and listening together.

No purchase or fabrication changes are made by this report. It consolidates existing analysis; it does not constitute a new acoustic validation or unconditional manufacturing release.

15 · Short glossary

16 · Sources and included evidence

This is a consolidation of saved project calculations, not a new measurement campaign. Manufacturer data and the crossover schematic are the model inputs; theoretical papers support modelling methods.

• SB Acoustics Tifa 6
• SB16PFCR25-4 woofer data
• SB26ST-C000-5 tweeter data
• Official Tifa 6 crossover schematic
• Miki porous-material model, 1990

The companion archive includes this self-contained report, current result files, historical B results, analysis code and electronics verification. Its manifest records file hashes. Original design and fabrication files remain in the C1 project package.