Centrifugal compressors: operation, reliability, start-up and maintenance

AMIN ALMASI is a lead mechanical engineer in Australia. He is chartered professional engineer of Engineers Australia (MIEAust CPEng – Mechanical) and IMechE (CEng MIMechE) in addition to a M.Sc. and B.Sc. in mechanical engineering and RPEQ (Registered Pro

INTRODUCTION

There have been many types of compressors such as reciprocating, screw, axial, etc; however, centrifugal compressors are known as main compressor options for many modern industrial and manufacturing plants. In past three decades, centrifugal compressors have replaced many other compressor types in many different services and applications. They are now most widely used compressor type in modern plants.

In general, centrifugal compressors have high technical demands and require elaborate reliability knowledge, maintenance and operational knowhow. The safe, reliable, and sustained operation of centrifugal compressors represent a key component for maximizing on-stream production in many plants. Reliability, operation and maintenance of centrifugal compressor trains are discussed.

CENTRIFUGAL COMPRESSORS

Centrifugal compressors have been used in many different applications from light gases to heavy gases and in a wide range of operational temperatures from cryogenic to hot gases. Most challenging ones are those with high head per impeller. Operators have usually tended to limit the head per impeller within certain limits for better operability, operational flexibility and reliability. As a very rough indication, 2500 – 4500 m per impeller might be noted as such limit depending on details of each service. As an average figure, 4000 m per impeller might be considered. For example, for a fluid with density of around 1 kg/m3, this means 0.4 – 0.5 Bar per each impeller. When gases compressed, the density is increased and higher pressure per impeller is expected. As another rough indication, for a fluid with density of around 10 kg/m3, the abovementioned means 3- 4 Bar per each impeller. Obviously for relatively heavy weight gases at high pressures (say above 80 Bar), the density is relatively high, say around or above 50 kg/m3, and high pressure rise by impeller (above 20 Bar per impeller) might be obtained for 3000 – 4000 m per impeller.

There have been compressors which can achieve a higher head per impeller; for example, even 6500 m or 7000 m per impeller or more. High head per impeller can be attained with the combination of high tip speeds, proper design using high strength materials, and high head coefficient (3D sophisticated impellers). These extremely high head impellers are not popular because they offer narrow operating ranges, risks of surge and operational problems. Relatively high head per impeller requires a high rotational speed, operation at high Mach numbers, a limited operating range, reduced efficiency, relatively low head rise-to-surge and difficulties in the control and operation. An optimization should be done as lower head per impeller means more impellers resulting in a long bearing span and potentials of rotordynamics issues; also, this is more expensive and results in heaver and larger compressors. Usually, any compressor for above 4500 m head per impeller should be dealt with a great care. Compressor’s efficiency is a critical matter. For most modern centrifugal compressors, the efficiency falls in the efficiency range of 80 – 88 percent. As more impellers are added to a compressor, operating ranges get effected and reduced; also, the combined efficiency, overall reliability and stability inherently might be reduced. The stability reduction is critical which is often related to dynamics, surge, stall, etc. One reason is there could be interactions between impellers; also, where there are many impellers, overall, the operating range of the assembly can be limited and there is more chance that any deviation or operating point movements lead to surge or other instability.

Map of each individual impeller affects the final map and the overall surge points; therefore, the final operating map and margin to the surge of a combined impeller assembly could be limited where there are many impellers in a compressor. Careful staging can allow taking advantage of the best operating range of each impeller, as there have been many successful compressors with many impellers. However, where there are many impellers, each can limit the final operating map of the compressor to some extent; therefore, the final map may be restricted. Each impeller should be carefully selected in concert with other impellers. Interactions between different compressor impellers and stages should be considered and matched. Overall, any compressor should have an acceptable operating range/map; preferably the range should be relatively wide range, considering rated condition(s) and alternative operating cases.

Another important operational consideration is the gas velocities in different part of a compressor. Compressor nozzle velocities should be carefully evaluated, and the nozzle should be properly sized. Nozzle velocities in 20 - 40 m/s range are commonly used depending on different factors such as gas molecular weight, etc. The higher the molecular weight and the lower the temperature, a lower velocity limit should be employed. For special situations such as revamp/renovation projects, relatively high velocity limits (even 40 m/s or more) might be accepted after thorough verifications. Relatively high nozzle velocities can translate into high pressure drops, operational problems, lower operating range, risks of some vibrations and more noise.

START-UP OF A COMPRESSOR

The compressor load at the start-up should usually be kept as low as possible to keep the starting torque at minimum possible. The low pressure of gases at the compressor start-up was often achieved by depressurizing the whole compression system; this is known as the depressurized star-up which was a very traditional and old-fashioned way of operation. Compressors then can be loaded once the driver reaches a suitable continuous speed. The depressurized star-up has allowed reducing the size of driver or starting device. However, this start-up method is not popular

anymore because of environmental and operational issues. In a plant, this depressurised starting method could result in:

• The loss of gases to depressurise compression loops after any plant shutdown.

• Flaring or destructions of a large amount of gases.

• Increased plant downtime; sometimes 5 – 12 hours required to complete plant start-up for a depressurised start-up.

As a consequence, the pressurized starting capability can help increasing the overall plant availability, performance, operation and efficiency. The incentives of pressurized starting capabilities led to significant development efforts for centrifugal compressors. For a compressor train, the driver sizing, design and operational details should be in a way to provide the capabilities for the pressurised start-up. Since the required start-up torque is mainly a function of compressor fluid load (pressure, etc) and rotating inertia, it is critical to fully understand the complete compression system as well as the whole driver assembly for a proper start-up arrangement. A rigorous dynamic simulation should also be performed to take into account start-up transients; this needs an accurate dynamic model and data, the actual control software and realistic operational performances. The pressurised starting is a modern requirement and should be considered for modern centrifugal compressors.

OPERATION, RELIABILITY SHUTDOWN, AND SURGE

An accurate dynamic simulation is needed in the different stages of selection and operation, for instance the support of sizing of many specific key elements in the compression system, commissioning, and reliability investigations. Generally, dynamic reliability simulations are fundamental to understand the operation of a compressor system in all operating situations especially transient cases. Particularly such an accurate dynamic simulation is critical for the operation as all important transient cases such as all cases of start-up, emergency shutdown, malfunction, etc should be properly simulated and required actions from operators should clearly be defined. Those elements which need help for sizing and verifications could be relief valves, control valves, suction/discharge valves, equipment sizing, timing parameters and others.

Among various operating scenarios, a compressor emergency shutdown (or compressor emergency trip) can represent one of high-risk cases which can expose the compressor to a severe surge condition. Major plant upsets can cause unwanted emergency shutdowns of compressor systems because there is no time for a normal shutdown procedure; such a shutdown imposes potential risks of surge on a compressor. An anti-surge system for any centrifugal compressor should be provided to protect the compressor from surge during an emergency shutdown. The investigation of histories of some plants has indicated that an emergency shutdown represents the highest risk to centrifugal compressors. In some cases, based on dynamic simulation results, a hot-gas-bypass anti-surge valve should be considered to avoid surge during such an emergency trip. Although such an anti-surge valve, which is in addition of normal anti-surge valve, should be considered as the last solution. Typically, such a hot-gas-bypass anti-surge valve is a fast-acting valve, to open in less than one second, to relieve the gas to the suction side and avoid surge at emergency shutdown. The route of such valve is a short distance between the discharge and suction and because there is no cooling on such a short loop, it is known as hot-gasbypass; it can only be used for a short time of shutdown.

Generally, anti-surge and control instruments should be as close as possible to the compressor nozzles. Common recommendation is 1 – 5 metres from the compressor nozzles for pressure and temperature measurement instruments. The flow measurement element which is a key element of any anti-surge system should usually be within 4 - 12 metres from the compressor nozzle. Too often, a flow measurement element is more of a challenge, given the additional requirements for straight runs of commonly-used flow elements; for example, “10 D” upstream and “5 D” downstream (D: diameter of the piping). Mounting instruments and sensors too far from compressor nozzles might lead to some operational issues such as slower action on changing variables and therefore slower reaction of the anti-surge system, which could result in some instability and damages.

MACHINERY SHOP AND SITE TESTS

Compressor performance testing at manufacturer’s shop and also at site are important requirements for nearly any centrifugal compressor train. The testing capability of compressors’ manufacturers should be confirmed during a bidding stage. It is necessary to verify the performance, mechanical running, rotordynamics, torsional behaviours and stability of operation of any compressor before its delivery to the job site.

The head coefficient and efficiency of an advanced centrifugal compressor varies with many specific parameters and operational details such as Mach number, etc. Therefore, such coefficients, parameters and details are usually unique in each application. On this basis, a performance testing is required to be conducted at the site rated conditions or conditions close to rated conditions. This is specifically true for the Mach number; therefore, the test should be at the rated Mach number or very close to it. Reynolds number is a dimensionless quantity that is used to help predict similar flow patterns in different fluid flow situations.

For compressors, like many other machineries, Reynolds number closely relates to the boundary layer and frictional losses. These characteristics should be evaluated very carefully for compressors. Only ASME PTC type-1 test can offer an accurate and useful performance test arrangement for centrifugal compressors.

Another important issue is correct prediction of surge points during a shop performance test. Surge points could be different if the piping and inter-stage facilities are varied. In other surge points are difficult to be reconstructed in the shop test. A smart decision could be using additional vessels or piping spools in the test arrangement, to simulate the discharge and suction gas volumes as close as possible to the site arrangement. For a variable speed centrifugal compressor, the surge points at different speed should be measured.

A considerable portion of a VSD electric driven compressor string testing is usually dedicated to the final validation of the VSD system and its integration with the main equipment as well as with the electrical power system. Shaft torque ripple is a well-known weak point of variable frequency systems; this ripple and the amount of harmonic disturbances injected by a VSD system into the surrounding electrical grid should closely be monitored with an advanced recording system. The general expectation is that the test result should validate low harmonic effects and low torque ripple, well within the rated values, therefore not requiring any mitigation measure such as an external suppression, additional harmonic filters, etc.

There is always concern about the vibration behaviour of a machinery train in a performance test. In addition to the job’s vibration monitoring systems including bearing vibration monitoring systems, casing vibration systems and others (which is provided for the compressor and shipped by the compressor), a thorough temporary vibration monitoring system should be installed on the piping, casings, bearing pedestals and support structure to get a better insight to all vibration and dynamic effects generated. This provides a complete picture of the train rotordynamics, overall machinery dynamic behaviours and whole structural vibrations.

AMIN ALMASI.