Old-fashioned high-pressure reciprocating compressors and oil-sealed heavy-bolted-head high-pressure centrifugal compressors have been used in high-pressure pipeline stations for decades. However, they seem unreliable, inefficient, heavy, bulky and wasteful. Modern high-pressure gas stations use vertically split, dry-gas-sealed, shear-ring head centrifugal compressors for high-pressure service.
Different design, manufacturing, performance and reliability aspects of modern high-pressure turbo-compressor trains are discussed here. New high-pressure turbo-compressors are designed with special focus on critical areas such as rotordynamics, thermodynamic performance, potential excitations, rotating-stall, auxiliary systems, fabrication processes and dry-gas seals. High-pressure turbo compressors will play more important roles in high-pressure gas pipelines in the future.
A vertically split casing with a heavy-bolted head is a classic design for high-pressure applications, which is not practical for very high pressures, because there is not enough space on the compressor head for the heavy bolts required. The commonly used configuration for modern high-pressure applications is a shear-ring cover head that eliminates the bolts at the compressor head. The end covers are retained by the shear-ring segments. The casing is generally forged from a suitable grade of steel.
High-pressure compressors generally use closed-type three-dimensional impellers. Both the efficiency and the pressure coefficients are higher for three-dimensional impellers. Impellers are usually manufactured from low-alloy steels. In-depth investigations of the material characteristics should be carried out to examine the material integrity with regard to various failure modes (such as sulfide stress cracking (SSC)). One of the major problems is that even after the transition of the gas to the dense phase, water can drop out if the gas is sufficiently cooled. For instance, SSC can occur in minutes if a susceptible material under sufficient stress is exposed to a wet sour condition in any wet upset (for example, for special sour gas pipelines).
For high-pressure compressors, the thrust force generated in the impeller assembly is much greater than can be handled in a thrust bearing. Therefore, the thrust forces must be balanced by using double entry impeller series (a back-to-back arrangement) and balance piston(s). A recommended design is a back-to-back impeller arrangement with two balance pistons. The remaining axial force is the difference between the forces which are a magnitude larger and an advanced double-acting tilting-pad thrust bearing should be used.
The shaft end seals in modern high-pressure compressors are generally dry-gas seals. Oil seals were common in older designs but are no longer used because of high power losses, low reliability, oil contamination and many other operational problems. The high pressure presents a challenge to dry-gas seal applications. There are a few dry-gas seal sub-suppliers for high-pressure services. Special attention is required because of the possibility of damage upon rapid depressurization and explosive decompression.
Figure 2: Internals of a modern gas turbine used for a high pressure pipeline compressor station application.
Electric motor drivers are used in small- and medium-size compression trains. One successful design is a two-end electric motor arrangement using a dedicated gear unit at each end to drive two compressor casings. Using this configuration, an optimum speed can be achieved for each compressor casing. The compressor bundles can easily be removed from each side.
For high-pressure compressor trains of more than 5 MW, gas turbine drivers have traditionally been used. Gas turbine drivers can be employed for compressor trains up to 100 MW.
Dynamic And Special Considerations
Vibrations at high-pressure turbo-compressors are usually controlled by employing very stiff rotor designs. The sufficient stability of the rotor should be ensured under any anticipated operating or malfunction conditions. Maximizing the rotor stiffness is achieved by limiting the number of impellers per casing. The best recommendations include the shortest possible bearing span and the maximum shaft diameter.
Modern high-pressure compressors (which are generally high-speed machines) only use tilting-pad bearings (or sometimes magnetic bearings). These bearings inhibit bearing cross-coupling excitation forces and instabilities.
In high-pressure applications, the fluid density could become high, which presents a great challenge for the design and dynamics of the machine. Above 300 barg, an average gas density exceeding 200 kg/m3 has been reported. This high fluid density lowers the rotating assembly’s natural frequency, which increases the risk of resonance (potential instability). For example, in a case study for a high-pressure gas centrifugal compressor, the natural frequency at the operational pressure was decreased to around 75% compared to the low-pressure operation. To prevent instability, the operating speed range should often be limited. The outlet width to passage width ratio of an impeller is an important factor in this regard.
An accurate calculation of gas properties is very important for the realistic modeling of a high-pressure compressor. The high density of the fluid increases the tendency to whirl. This high density could magnify the excitation forces generated in balance piston seals and other internal seals, which can result in instability. The stiffness and damping properties of the various seals within a compressor are strictly related to the seal’s shape, gap, design, manufacturing and quality control.
Special attention is required for seal selection, design and inspection. Generally, various methods are used to eliminate the magnitude and effect of cross-coupling forces. Features that inhibit the tangential swirl can help to reduce the cross-coupling forces. Such special features may only be needed above pressures of 130 barg. These solutions are usually based on empirical data, comparison with successful designs and testing at the full pressure.
A “shaker test” is usually required for a high-pressure compressor to verify the rotordynamics behavior and stable operation. The main aim of this test is to determine the damping of the rotor system at various operating conditions spread over the entire range of compressor speeds and characteristics. Generally, a suitable broad-range vibration generation device (a shaker) should be employed. Based on a modern procedure for performing this test, the end cover of the barrel carrying the rotor is fitted with a temporary extension designed to host an active magnetic bearing at the rotor shaft extension. This magnetic bearing is then operated in a manner that induces vibrations on the rotating shaft.
The overall damping of the compressor is determined by evaluating the rotor vibration response to the non-synchronous excitations produced by the exciter. During the test, usually around 12–20 operating points are measured and the modal frequency and damping are evaluated. Generally, the compressor should reveal a very satisfying stability (super-critical damping) behavior and a very low vibration level under all operating conditions and excitation modes. A limited number of compressor manufacturers offer the “shaker test” for verifying a compressor’s dynamic performance. The requirements should be fixed before the compressor order.
Specific attention should be given to the rated head to include a suitable margin covering gas property uncertainties, fouling and performance deterioration in high-pressure services.
The material selection, stress analysis, fatigue study and quality control of compressor shafts are important. Sufficiently low stresses (as low as around 20% of the minimum yield strength) are recommended for regions of the shaft that might be wetted by the gas. Impeller cracking caused by welded fabrication is a widespread problem. The electro-discharge machining process (instead of welding) is recommended for the fabrication of high-pressure impellers. Modern finite element (FE) studies using fine meshing techniques are required, with a special focus on highly stressed localized areas.
The best vendor shop test package for high-pressure centrifugal compressors is a pressurized shop mechanical run test with nitrogen followed by a shop performance test as per ASME PTC 10 type 1 with the job gas. Some vendors offer ASME PTC 10 type 2 with inert gases (such as CO2, N2, helium, or similar) or their mixtures to simulate the gas. This is not recommended because ASME PTC 10 type 1 with the job gas better simulates the site conditions.
A shop performance test is more useful if the real gas is used and the pressures, compressor speed, capacity and power match the site’s operating conditions as close as possible. Such a performance test can provide very valuable information about the likely behavior of a high-pressure compressor such as certain aspects of the aerodynamic excitations, operation near the limits (such as the surge limit or the choke), instability and realistic effects of the seal or bearing. These effects are of concern in high-pressure compressors, even for those from the best-known vendors.
Small pressure fluctuations caused by instability (such as a surge or a stall) can be observed using special pressure measuring instruments. Based on shop performance test results, some modifications to a compressor’s operating range may be required. Often, the rotating-stall imposes some limitations on the minimum speed (the minimum flow without the recycling).
Amin Almasi is lead rotating equipment engineer at WorleyParsons Services Pty Ltd, Brisbane, Australia. He can be reached at email@example.com.