Pressure Pulsation ใน Screw Blower: สาเหตุและวิธีลดทอน

As AERZEN has engineered rotary positive displacement machines since 1864, pressure pulsation management in screw blower systems remains one of the more nuanced challenges that process engineers encounter during system commissioning and long-term operation. This article addresses the physics of pulsation generation, its downstream effects, and the engineering approaches used to design effective damping systems.

หมายเหตุ: ตัวเลขในส่วนนี้เป็นตัวอย่างสมมติเพื่ออธิบายหลักการ — Case data and numerical illustrations in this article are ตัวอย่างสมมติ prepared for educational reference. Figures reflect industry-typical ranges.

What Causes Pressure Pulsation in a Screw Blower?

The Fundamental Mechanism

A screw blower — such as AERZEN’s Delta Screw series — compresses gas by trapping a volume of air between two counter-rotating helical rotors and progressively reducing that volume as the rotors turn. Unlike a centrifugal fan, which delivers a near-continuous flow, a positive displacement screw blower releases compressed gas in discrete pulses corresponding to each rotor lobe pass.

The primary pulsation frequency, known as the blade passing frequency (BPF), is determined by:

BPF (Hz) = (Rotational speed, RPM / 60) × Number of lobes per rotor

For example, a screw blower running at 3,000 RPM with a 4-lobe male rotor produces a fundamental BPF of:

(3,000 / 60) × 4 = 200 Hz

Harmonics at 400 Hz, 600 Hz and above are also generated, each with diminishing amplitude in a well-designed machine.

Why Pulsation Amplitude Matters

Pressure pulsation amplitude is typically expressed as a percentage of mean discharge pressure (ΔP/P_mean × 100%). In a properly designed screw blower with internal compression — that is, a machine whose built-in pressure ratio approximates the system pressure ratio — pulsation amplitude at the discharge port can typically be held below 2–5% of mean discharge pressure.

When the system pressure deviates significantly from the machine’s built-in pressure ratio, under-compression or over-compression occurs at the discharge port, causing a sharp pressure step as the rotor pocket opens to the discharge line. This step excites broadband acoustic energy in addition to the BPF tones and substantially increases pulsation amplitude.

Downstream Effects of Unattenuated Pulsation

Piping Vibration and Fatigue

Pressure pulsations propagate as acoustic waves through the connected piping system. When the acoustic natural frequency of a pipe section or branch coincides with a pulsation frequency (or its harmonic), standing waves form and pipe wall vibration amplitudes increase significantly. Sustained vibration at these levels causes fatigue at welded joints, tee connections, and instrument tap locations.

The evaluation of mechanical vibration on rotating machines and their supporting structures is addressed in ISO 10816-3:2009 Mechanical vibration — Evaluation of machine vibration by measurements on non-rotating parts — Part 3: Industrial machines with nominal power above 15 kW and nominal speeds between 120 r/min and 15 000 r/min — https://www.iso.org/ (verified 2026-05-08)

While ISO 10816-3 addresses vibration at the machine body, the same principles of acceptable vibration velocity (mm/s RMS) apply as acceptance criteria for piping vibration assessments in adjacent systems.

Instrument and Control System Interference

Pressure transmitters, flow meters (particularly Coriolis and differential pressure types), and control valves are susceptible to measurement error and accelerated wear when exposed to high-amplitude pulsation. A pressure transmitter with a diaphragm resonance frequency near 200 Hz in a system with a 200 Hz BPF may report erratic readings, causing unnecessary control interventions.

Process Performance Degradation

In pneumatic conveying applications, high pulsation amplitude causes intermittent flow velocity variation that results in material settling in horizontal conveying lines and increased degradation of fragile bulk solids. In aeration applications, pulsating airflow produces uneven distribution across diffuser arrays, reducing oxygenation efficiency.

Acoustic Modeling: The Engineering Foundation of Damping Design

Transfer Matrix Method

The industry-standard analytical approach to pulsation damping design is the Transfer Matrix Method (TMM), also known as the four-pole method. The piping system is modeled as a series of acoustic elements — pipe sections, volume chambers, expansion chambers, and branches — each represented by a 2×2 transfer matrix relating acoustic pressure and volume velocity at inlet and outlet.

By cascading the element matrices, the overall acoustic response of the system can be computed, and the frequencies at which acoustic resonances (standing waves) occur can be predicted before physical installation.

Reference methodology is described in: Wiley — Compressor Handbook (Paul C. Hanlon, ed.), Chapter on acoustic analysis of pulsating flow systems. For standardized analytical procedures for gas pulsation in reciprocating and rotary compressor piping, refer to API Standard 618: Reciprocating Compressors for Petroleum, Chemical, and Gas Industry Services (5th ed.) — https://www.api.org/products-and-services/standards/important-standards-announcements/standard-618 (verified 2026-05-08). While API 618 addresses reciprocating machines directly, its acoustic simulation methodology (Chapter 3 design approach) is routinely applied to rotary positive displacement blower piping analysis.

Finite Element Acoustic Modeling

For complex 3D piping geometries with multiple branches, tee junctions, and inline components, Finite Element (FE) acoustic modeling provides higher accuracy than analytical TMM. Software such as ANSYS Mechanical Acoustics or COMSOL Multiphysics Acoustics module can solve the Helmholtz equation across the full piping network geometry, identifying pressure nodes and anti-nodes throughout the system.

This approach is particularly valuable when:

• Piping routing cannot be changed (constrained plant layout)
• Multiple blowers discharge into a common header
• Acoustic interaction between pulsation sources must be assessed

Damping Design Strategies

1. Suction and Discharge Pulsation Dampeners (Bottle Silencers)

Volume bottle silencers are the most commonly applied passive pulsation control device. An expansion chamber with a volume typically 10–20 times the swept volume of the blower per revolution is installed immediately at the suction or discharge connection. The abrupt area expansion causes acoustic energy to be partially reflected back toward the source, reducing transmitted pulsation amplitude at the blade passing frequency.

Design parameters:

• Chamber volume (V_chamber): target volume ratio V_chamber / V_displacement ≥ 10–20 (higher ratio = broader attenuation bandwidth)
• Chamber length-to-diameter ratio: typically 3:1 to 5:1 for broadband attenuation
• Inlet/outlet nozzle sizing: nozzle area contraction ratio affects low-frequency performance

2. Helmholtz Resonators (Tuned Silencers)

A Helmholtz resonator is a side-branch cavity connected to the main pipe via a short neck. When tuned to a specific excitation frequency, it acts as a dynamic absorber, presenting high acoustic impedance at that frequency and dramatically reducing transmitted pulsation amplitude.

The resonant frequency of a Helmholtz resonator is:

f_H (Hz) = (c / 2π) × √(A_neck / (V_cavity × L_eff_neck))

Where:

• c = speed of sound in the gas (m/s)
• A_neck = cross-sectional area of the neck opening (m²)
• V_cavity = volume of the resonator cavity (m³)
• L_eff_neck = effective acoustic length of the neck, including end correction (m)

At 20°C dry air: c ≈ 343 m/s. For process gas at elevated temperature, c scales as √(T_abs / 293), so the resonant frequency shifts accordingly. This temperature dependence must be accounted for if the resonator is designed at ambient conditions but operates at elevated discharge temperatures.

Helmholtz resonators deliver high attenuation at a narrow frequency band. They are most effective when a single dominant BPF tone causes the majority of the pulsation problem.

3. Expansion-Chamber / Quarter-Wave Silencers

A quarter-wave stub — a closed side branch with length equal to one quarter of the acoustic wavelength at the target frequency — presents an acoustic short circuit at its resonant frequency, effectively reflecting pulsation energy back to the source.

L_stub (m) = c / (4 × f_target)

Example: To attenuate 200 Hz at 40°C (c ≈ 355 m/s): L_stub = 355 / (4 × 200) = 0.444 m

Quarter-wave stubs are simple to fabricate and highly effective for single-frequency targets, but their performance degrades rapidly off the design frequency.

4. Flexible Hose and Pipe Expansion Joints

Flexible connections with appropriate acoustic compliance interrupt the structural transmission path between the blower casing and the connected piping. They are particularly effective at reducing structure-borne vibration at high frequencies. However, flexible elements do not provide significant acoustic attenuation for low-frequency pressure wave propagation through the gas column and should not be the sole mitigation strategy for low-frequency pulsation.

5. Variable Speed Drive (VSD) — Frequency Avoidance

When a screw blower is equipped with a VSD — as in AERZEN’s Delta Screw E-Compressor or variable-speed BVS configurations — the rotational speed can be controlled to avoid operating at speeds where the BPF aligns with a known piping acoustic resonance. This approach requires prior knowledge of the system’s acoustic natural frequencies from modeling or measurement, and active control logic to maintain exclusion zones in the speed map.

Field Measurement and Validation

After installation of damping devices, field validation should confirm that pulsation amplitude has been reduced to acceptable levels. The recommended measurement approach:

1. Install dynamic pressure transducers (fast-response, minimum 10× BPF bandwidth) at defined measurement points: blower discharge nozzle, upstream and downstream of each silencer, and at the process connection.
2. Acquire time-domain pressure signals and convert to frequency domain via Fast Fourier Transform (FFT).
3. Compare measured BPF amplitude against the design acceptance criterion (typically ΔP/P_mean < 2% at BPF, per API 618 Approach 3 guidance adapted for rotary machines).
4. Evaluate overall vibration velocity (mm/s RMS) at adjacent piping per ISO 10816-3 Zone A/B limits for the relevant machine class.

ตัวอย่างสมมติ Case: Pulsation Attenuation on a Delta Screw Blower — Aeration Header

A wastewater treatment facility (anonymized — ตัวอย่างสมมติ) installed a 75 kW screw blower on an aeration air header serving twelve membrane diffuser zones. During initial commissioning, a 4-inch common header exhibited visible pipe vibration and differential pressure transmitters reported ±8% pulsation amplitude at 3,000 RPM.

Acoustic modeling identified a standing wave resonance in a 14-metre header section at 186 Hz, close to the blower BPF of 200 Hz. Two corrective measures were implemented:

1. A volume bottle silencer (volume ratio 15:1 relative to displacement volume) was installed at the blower discharge connection.
2. VSD control was programmed to exclude the 2,850–3,150 RPM band from the operating range.

Post-modification measurement confirmed pulsation amplitude reduced to 1.9% at the header midpoint and pipe vibration velocity dropped to Zone A (acceptable) per ISO 10816-3. Diffuser distribution uniformity, measured by individual zone flow balancing, improved measurably.

Frequently Asked Questions

Q1: Is pressure pulsation from a screw blower worse than from a Roots blower? A Roots (rotary lobe) blower compresses without internal compression — the rotor pockets open directly to the discharge line at system pressure, creating a sharp pressure step and relatively high pulsation amplitude, particularly at low pressure ratios. A screw blower with appropriate built-in pressure ratio delivers gentler internal compression and lower pulsation amplitude for the same duty. The trade-off is mechanical complexity and cost.

Q2: Does gas composition affect damping design? Yes. The speed of sound varies with molecular weight and temperature. For heavier gases (CO₂, methane mixtures) c is lower, shifting all acoustic natural frequencies downward. Damping devices designed for air duty must be re-calculated for alternative gas service. Contact AERZEN engineering team when specifying blowers for non-air gases.

Q3: What is the typical insertion loss achievable with a well-designed bottle silencer? A properly sized expansion-chamber silencer can achieve 10–20 dB insertion loss at the blade passing frequency. Combined with a Helmholtz resonator or quarter-wave stub targeting a dominant harmonic, total attenuation can exceed 25 dB at the primary excitation frequency.

Q4: Can pulsation cause damage to the blower itself? If acoustic resonance in the discharge piping causes reflected pressure waves to arrive back at the blower discharge port in phase with the rotor opening event, the instantaneous discharge pressure can transiently exceed the operating design pressure. In severe cases this causes rotor-to-housing contact, premature bearing wear, and seal damage. Early-stage monitoring detects this before it escalates.

Q5: What documentation should I request from the blower supplier for acoustic analysis? Request: (a) discharge pressure pulsation spectrum data from OEM testing at rated conditions, (b) swept volume per revolution, (c) rotor lobe count and rotational speed range, and (d) any pulsation testing reports. AERZEN provides full technical data packages for machines supplied under rental and Subscription Plan agreements.

Contact AERZEN Engineering Team

For pulsation analysis consultation, technical data packages, or discussion of damping system design for your application:

AERZEN Rental Thailand Website: www.aerzenrentalth.com Office: 038-015-488 Hotline 24/7: 098-323-2626 Email: thai@aerzenrental.com

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