In the context of increasing global energy demands and the shift towards cleaner fuels, liquefied natural gas (LNG) has emerged as a pivotal medium for transportation, storage, and reception of natural gas. The reliability of LNG processes and equipment is crucial for regional energy security. Among various LNG liquefaction processes, the mixed refrigerant liquefaction process dominates due to its advantages in base-load and peak-shaving applications. The heart of this process is the liquefaction unit, where the refrigerant compressor serves as a key component. Traditionally, steam turbines have been used to drive compressors in large-capacity LNG systems. However, with the maturation of electric drive system technology, there is a growing trend towards “steam-to-electric” conversion. Electric drive systems offer benefits such as shorter start-up times, stable operation, ease of use, lower maintenance costs, and simplified repair procedures. Despite these advantages, the development of low-temperature, large-capacity LNG platform electric drive systems remains challenging, with products primarily supplied by renowned international companies. This paper, based on the domestic development of a refrigerant compression electric drive system for a large LNG project, presents a comprehensive analysis and research on the structural aspects of such systems from a first-person perspective.
The electric drive system is integral to providing power for the entire LNG process. It typically consists of a motor and a frequency converter, driving the compressor via a gearbox for speed increase. In this project, the motor system includes a large synchronous motor and an excitation rectification device. The design must ensure safe and stable operation under high parameters and harsh environmental conditions. Key challenges include low-temperature environments (as low as -52°C), high capacity and speed, wide speed regulation ranges, stringent vibration and noise requirements, and the need for vibration isolation foundations. Additionally, factors like lifting, maritime transport, and platform installation under extreme loads must be considered.
To address these challenges, I conducted detailed analyses on the mechanical characteristics of the electric drive system, focusing on structural strength, rotor dynamics, and vibration characteristics of the spring isolation foundation. The electric drive system’s performance hinges on robust design and validation through simulation and testing. In the following sections, I will outline the main technical parameters, delve into the structural analyses, and compare simulation results with experimental data.
Main Technical Parameters of the Electric Drive System
The electric drive system is designed to meet the specific requirements of the LNG project. Below are the key technical parameters of the synchronous motor, which form the basis for structural analysis.
| Parameter | Value |
|---|---|
| Model | TZW80000-2 |
| Rated Power | 80 MW |
| Rated Voltage | 10 kV |
| Number of Poles | 2 |
| Efficiency | 98.85% |
| Power Factor | 1.0 (leading) |
| Rated Current | 4,450 A |
| Rated Frequency | 60 Hz |
| Rated Speed | 3,600 r/min |
| Speed Regulation Range | 2,400 to 3,960 r/min |
| No-load Excitation Current | 280 A |
| Rated Excitation Current | 620 A |
| Rated Excitation Voltage (at 110°C) | 219 V |
| Short-Circuit Ratio | 0.52 |
| Ambient Temperature | -52°C to +30.1°C |
| Vibration | 2.3 mm/s (bearing housing), 65 μm (shaft vibration) |
| No-load Noise | 85 dB(A) |
| Protection Class | IP55 |
| Cooling Method | IC86W |
| Bearing Type | Sliding bearings |
| Explosion-proof Class | Exp IIA T3 Gc |
These parameters underscore the high-performance demands on the electric drive system, necessitating rigorous structural analysis to ensure reliability under operational conditions.
Structural Analysis of the Electric Drive System
The structural integrity of the electric drive system is paramount for its functionality. I focused on three core areas: rotor strength, rotor dynamics, and spring isolation foundation vibration characteristics. Each aspect was analyzed using finite element methods (FEM) and computational tools to optimize design and validate performance.
Rotor Strength Analysis
The rotor is a complex assembly comprising the shaft, coils, retaining rings, excitation devices, and rectifier discs. Due to high operational speeds (up to 4,356 r/min overspeed) and large diameters, centrifugal forces induce significant stresses. The retaining ring assembly is particularly critical, as it must withstand centrifugal forces from end windings and itself, along with stresses from interference fits with the shaft and center ring.
I employed ANSYS software for a 3D finite element analysis using a half-tooth, half-slot model. Contact pairs were defined with friction, using CONTA174 and TARGE170 elements. Interference fits were incorporated geometrically. The analysis aimed to ensure that stresses in all components remained within allowable limits under static, rated, and overspeed conditions. The equivalent stress results are summarized below.
| Component | Static Condition (MPa) | Rated Condition (MPa) | Overspeed Condition (MPa) | Design Requirement (MPa) |
|---|---|---|---|---|
| Retaining Ring | 734 | 660 | 845 | < 900 |
| Shaft | 723 | 575 | 745 | < 750 |
| Center Ring | 380 | 183 | 171 | < 760 |
| Slot Wedge | 512 | 333 | 365 | < 520 |
The stress distributions confirmed that all components met strength criteria. Additionally, contact pressure analyses ensured that interference fits remained effective under rated conditions, providing sufficient constraint for alignment. The governing equation for centrifugal stress in rotating components can be expressed as:
$$ \sigma_c = \rho \omega^2 r^2 $$
where $\sigma_c$ is the centrifugal stress, $\rho$ is material density, $\omega$ is angular velocity, and $r$ is radius. This highlights the importance of speed and geometry in design.
Rotor Dynamics Analysis
Rotor dynamics analysis is essential to avoid resonance and ensure stable operation. The shaft system includes the motor rotor, coupling, gearbox shaft, and compressor shaft, with compressor speeds ranging from 3,899 to 5,849 r/min. I used MADYN2000 to model the system, focusing on lateral and torsional vibrations.
Lateral Vibration Analysis
The motor rotor is supported by two pedestal bearings with tilting pad bushings. Bearing parameters such as specific pressure and oil temperature were verified. Critical speeds were calculated to ensure separation from excitation frequencies. The first three critical speeds are listed below.
| Mode | Critical Speed (r/min) | Vibration Shape |
|---|---|---|
| 1 | 1,455 | First bending |
| 2 | 4,786 | Second bending |
| 3 | 5,187 | Bending at exciter end |
These critical speeds are outside ±20% of the motor’s operating range (2,400–3,960 r/min), meeting API 617 standards. Unbalance response analysis was performed, with shaft vibration at the excitation end bearing shown in the frequency domain. The peak amplitudes at critical speeds were within limits.
Torsional Vibration Analysis
Torsional natural frequencies must avoid excitation frequencies from motor and compressor operation. The shaft system’s torsional modes were analyzed, with results for the first eight frequencies provided.
| Mode | 1 | 2 | 3 | 4 | 5 | 6 | 7 | 8 |
|---|---|---|---|---|---|---|---|---|
| Frequency (Hz) | 16.9 | 34.5 | 238.4 | 256.3 | 308.8 | 346.7 | 444.9 | 485.5 |
The motor frequency range is 40–66 Hz, and the compressor frequency range is 64.98–97.48 Hz. All torsional frequencies are outside ±10% of 1x and 2x these ranges, ensuring safe operation. The torsional vibration equation for a multi-mass system is:
$$ J \ddot{\theta} + C \dot{\theta} + K \theta = T(t) $$
where $J$ is inertia matrix, $C$ is damping matrix, $K$ is stiffness matrix, $\theta$ is angular displacement, and $T$ is torque.
Spring Isolation Foundation Vibration Characteristics Analysis
The electric drive system is mounted on a spring isolation foundation to mitigate vibrations. Based on EN 1993 standards, I designed a foundation with 10 spring units, each with vertical stiffness $K_v = 16,651.1 \, \text{N/mm}$ and horizontal stiffness $K_h = 12,121 \, \text{N/mm}$. A finite element model was built to analyze modal and harmonic responses.
Modal analysis yielded the first three natural frequencies and modes:
| Mode | Natural Frequency (Hz) | Participation Factors (X, Y, Z) | Mode Shape |
|---|---|---|---|
| 1 | 3.79 | (0.00, 0.00, 0.94) | Vertical translation |
| 2 | 3.85 | (0.77, 0.01, 0.00) | Horizontal translation in X |
| 3 | 5.43 | (0.02, 0.96, 0.00) | Horizontal translation in Y |
Harmonic response analysis under unbalance loads (balance grade G2.5) was conducted. Vibration responses at bearing locations were evaluated. The maximum vibration velocities in the operating range are summarized below.
| Location | Direction | Maximum Vibration Velocity (mm/s) | Frequency at Peak (Hz) | Limit (mm/s) |
|---|---|---|---|---|
| Bearing 1 | Vertical | 0.44 | 53.04 | 3.5 |
| Bearing 1 | Horizontal | 1.63 | 52.68 | |
| Bearing 2 | Vertical | 1.24 | 59.96 | |
| Bearing 2 | Horizontal | 1.80 | 59.96 |
All values are within the 3.5 mm/s limit, confirming the foundation’s effectiveness. The vibration transmission can be modeled using a spring-mass-damper system:
$$ m \ddot{x} + c \dot{x} + k x = F_0 e^{i \omega t} $$
where $m$ is mass, $c$ is damping, $k$ is stiffness, and $F_0$ is excitation force.
Comparison of Simulation and Experimental Results
To validate the electric drive system design, I conducted full-scale dynamic balancing tests. Vibration at key points was monitored and recorded. After balancing, shaft vibration Bode plots at the excitation end bearing were obtained. A comparison between simulation and experimental results is presented below.
| Parameter | Simulation Result | Experimental Result | Difference |
|---|---|---|---|
| First Critical Speed | 1,455 r/min | 1,444 r/min | 1% |
| Shaft Vibration Amplitude at First Critical (X-direction, P-P) | 17.8 μm | 14.5 μm | 23% |
| Shaft Vibration Amplitude at First Critical (Y-direction, P-P) | 23.6 μm | 26.5 μm | 12.3% |
The close agreement between simulation and experiment demonstrates the accuracy of the analytical models and validates the structural integrity of the electric drive system. Discrepancies are within acceptable engineering tolerances, attributable to factors like material property variations and manufacturing tolerances.
Conclusion
In this research, I successfully completed the domestic development of a low-temperature, large-capacity LNG electric drive system, meeting all technical requirements for a large LNG project. The structural analyses—encompassing rotor strength, rotor dynamics, and spring isolation foundation vibration—provided critical insights and optimization strategies. The use of finite element methods and dynamic simulation tools enabled precise design validation, ensuring reliability under extreme conditions. The comparison with experimental data confirmed the robustness of the approach. This work not only supports the deployment of advanced electric drive systems in LNG applications but also serves as a reference for future projects involving high-performance rotating machinery. The electric drive system’s role in enhancing operational efficiency and safety underscores its importance in modern energy infrastructure.
Throughout this study, the electric drive system has been central to addressing the challenges of LNG liquefaction. By integrating advanced structural analysis with practical testing, I have demonstrated that domestically developed electric drive systems can achieve international standards. Future work may explore further optimizations in materials, cooling systems, and control strategies to push the boundaries of electric drive system performance in cryogenic environments.