Amin Almasi is senior rotating equipment consultant in Australia. He is chartered professional engineer from Engineers Australia and IMechE and registered professional engineer in Australia and Queensland (M.Sc. and B.Sc. in mechanical engineering). He specializes in rotating equipment, condition monitoring and reliability.

Amin Almasi

Centrifugal pumps are used in many plants and facilities. Many are operated in parallel. There are risks, difficulties, and concerns to consider for parallel operation of pumps. There are factors and parameters which should be considered to mitigate these risks.


Many pumps, after all, have been damaged due to poor configuration, wrong selection, or operational carelessness in parallel operation. Further general knowledge on pump parallel operation is relatively low. There has been limited publication material available on this important topic.


Unfortunately, some engineers have been led to believe that pumps are like home appliances. They think they can put them in parallel operation and expect trouble-free operation. This is not true. The fact that pumps in parallel operation take suction from a common source and discharge into a common header is often overlooked. The fact that the pumps interact with each other is also poorly understood.


Different types of pump, of course, present different behaviours when in the parallel operation. As an example, so called “sensitive pumps”, which are more affected by operation far from BEP (best efficiency point), can be more troublesome when used in parallel operation.

There have been many reasons to put pumps in parallel operation. Sometimes, it is to cover a wide range of flow requirements. In other situations, they are installed in parallel for standby requirements or emergency backup. In an expansion or renovation program, a few new pumps may be added to an existing set to increase pumping capacity.

When pumps run in parallel, they would be expected to operate against the same discharge head. The combined pump “head-capacity” performance curve is, theoretically, determined by adding the respective flow rates of each pump at a series of specific head values. However, this is purely theoretical. There are many practical factors and parameters which should be considered.

If pumps are not properly selected for parallel operation, it can lead to inefficient operation, pump unreli ability, damage, and other problems.

If pumps are not properly selected for parallel operation and all required provisions for potentially complicated operation are not foreseen, it can lead to inefficient operation, pump unreliability, damage, and other operational problems. When wrongly selected pumps are operated in parallel, one of the pumps might be driven outside of its allowable operating range (far from BEP), even near shutoff point. This can result in overheating and pump damage.

Parallel operation of identical pumps is a better-known topic than operation of dissimilar pumps. There are established rules for the parallel operation of identical pumps. Performance curves should be matched within specified tolerances. In other words, although pumps are theoretically identical, because of differences in manufacturing and tolerances, their curves are slightly different. The differences between curves should be limited to 2% to avoid problems.


Dissimilar pumps operating in parallel introduce engineering complexity. The head should be matched at rated point of each pump (not at the same volumetric flow rate points). Parallel operation of dissimilar pumps is most often seen in revamp, renovation, or expansion projects.

For a revamp/renovation case, the head of new pump rated point should be matched with the head of the existing pump rated point (each has its own volumetric flow rate) considering the elevation head difference and frictional head difference (at downstream) for each pump. The surrounding area of rated points should offer a suitable operating condition for all pumps.

When dissimilar pumps are operated in parallel, however, there is risk if any pump is pushed to operate below its minimum allowable flow rate. As a practical measure, the pump manufacturer should be asked to simulate operation and explain all operational details and guarantee trouble-free operation for all combinations of parallel operation. In addition, operating points for all operating scenarios should be plotted and compared between the minimum continuous flow and the end of curve. A full simulation exercise and a study report for all possible parallel operation cases should be requested.


The shape of the “head–capacity” performance curve should be carefully considered when selecting pumps for parallel operation. If the pump curve droops (the head drops) as the flow is reduced towards shutoff point, there can be operational problems. In this case, when the first pump is started, a second pump may not be able to move from the shutoff point. This pump could run near shutoff, overheat, and possibly fail.

In other words, pumps in parallel operation work against the same discharge head. For a curve with more than one flow condition for a given head, a pump can be stuck near shutoff. This pump could fail over time. As a general rule, the performance curve should continuously rise to the shutoff point. In other words, pumps with head-capacity curves that droop towards shutoff point or have more than one flow condition for a given head, should not be operated in parallel. 

Amin Almasi is a Chartered Professional Engineer in Australia and U.K. (M.Sc. and B.Sc. in mechanical engineering). He is a senior consultant specializing in rotating equipment, condition monitoring and reliability

Articles

CHALLENGES AND DIFFICULTIES OF PARALLEL OPERATION 

TURBO TIPS: CENTRIFUGAL PUMPS

This article discusses how process and operation conditions can affect turbocompressors and their operation, performance, and reliability through three real-world case studies that address how downstream, upstream, the process itself, and all surrounding facilities and items impact turbocompressors.

Radial bearings of tilting-pad type in a critical centrifugal compressor showed spikes in bearing pad temperatures every 1 - 4 hours. Measured temperatures went from 96°C (baseline) to 115-119°C (spike). The recommendation by the operations team was to reduce the inlet lubrication oil temperature to limit the peak of the spikes below high temperature limits. The initial proposal was to reduce the temperature of the delive red lubrication oil to the compressor from 50°C to 45°C. It was implemented. However, the lubrication oil temperature reduction caused a higher temperature at the bearing pads. This might be because of more viscosity of lubrication oil and consequently more friction. In this case, a 5°C reduction in the lubrication oil supply temperature caused around 2-3°C increase in bearing (resistance temperature detector (RTD) temperatures. The reduction of supply lubrication oil temperature could not reduce temperature spikes at RTDs.


This is an example of how random actions cannot solve the problem. The root-cause of the problem should be identified and eliminated. After thorough studies, the root cause of temperature spikes was identified as varnish in the lubrication oil. Every time varnish settled on a bearing pad RTD, a spike in the bearing RTD temperature was recorded.


One indication that helped to correctly identify the root cause was a simple test of switching the lubrication oil filter to see if spike patterns changed. The switch of filter reduced varnish deposits on bearings temporarily and this reduced spikes in RTD bearing temperature and changed its patterns.

Investigation showed one of additives in the lubrication oil interacted with compressor process gas and produced varnish. It was decided to change the lubrication oil type for this compressor. The operation of the compressor has been smooth and satisfactory since this adjustment. 

The pressure ratio of a process centrifugal compressor was adjusted due to a change in gas composition and operating conditions. The pressure ratio (discharge pressure to suction pressure) could move theoretically from 2.25 to 2.51 because of these changes (gas composition change, higher required discharge pressure, etc). The compressor was initially rated for a pressure ratio of around 2.39.

Process and operation conditions can affect turbocompressors and their operation, performance, and reliability

The compressor data including compressor map, design conditions, compressor details, and thrust bearing data were carefully reviewed. Rises in differential pressure and load on thrust bearings were accurately determined and verified. Based on these studies, it was confirmed that there was no issue for the thrust bearing regarding this differential pressure rise. A thrust bearing should often be sized to absorb this kind of change. In this case, these changes were within thrust bearing limits. In addition, there were temperature measurements and vibration monitoring for the thrust bearings. These could highlight possible malfunctions. The compressor has been operated safely and reliably with the modified operating conditions.


The potential of liquid carryover to a refrigeration centrifugal compressor was raised by the operation team of a plant. Recorded data showed that prior to start-up of the compressor, the suction line was at 6.5 Barg and 5°C. Under these conditions, the refrigeration gas was theoretically liquid and could damage the compressor, seals, and inlet strainer when it started with liquid flowing through the centrifugal compressor. It was recommended by the operations team to depressurize the refrigeration gas to below 4.5 Barg prior to the start-up. In addition, they asked to investigate possible solutions to detect or automatically drain liquid refrigeration from the inlet piping at pre-start.


Investigations showed that liquid carryover was a false alarm. The temperature transmitter range was up to 5°C. It showed 5°C for any temperature above 5°C. The actual temperature in the suction line before the start-up was always above 20°C. Theoretically, there was a little chance for refrigeration gas conditions reaching the saturation point during a shutdown.


It was decided to change the temperature transmitter with a new one with a measurement range up to 40°C. There might be a remote chance that the gas condition during a shutdown could reach saturation point. In that case, draining of the collected liquid would be necessary. The operation procedure was corrected and a necessary step for the drain of liquid at the suction drain prior to the compressor start-up was included. ■

Amin Almasi is a Chartered Professional Engineer in Australia and U.K. (M.Sc. and B.Sc. in mechanical engineering). He is a senior consultant specializing in rotating equipment, condition monitoring and reliability.

How operation & process can affect turbocompressors

TURBO TIPS: CASE STUDIES

Fans have been widely used in many facilities and plants to move gases. Supply and exhaust fans have also been used in different systems to avoid excessive pressure build-up in the equipment or service being served. Forced and induced-draft fans, too, have been used to maintain a specified draft over the intended application. A fan should be selected to match its downstream and upstream systems to provide a proper flow of gas.


Although one operating fan and another standby configuration has been used, multiple parallel operation is discouraged. Many experts have recommended using just one operating fan at a time (one operating and another standby) on each system, avoiding parallel operating fans. This is to reduce complexity and operational problems as parallel operation has always been complicated and is usually problematic. On the other hand, there have been various reasons for using multiple fans in parallel operation on some special systems. In some specific cases, such as some revamp and renovation projects two fans may be provided to working in parallel. Capacity control by various fan combinations may be more economical than other control methods in some specific cases. Multistage fans may be necessary when pressure requirements exceed the capabilities of a single-stage fan.


When two fans are used in parallel, they may be located quite remote from each other or close together. Sometime two parallel side-by-side fans are close enough to share some items such as lubrication system, shaft, bearings, and casing. Double-width, double-inlet fans are essentially two fans in parallel in a common housing. However, the above-mentioned sharing of items or parts should be avoided as it reduces the overall reliability. Two completely independent fans have always been preferred.


Parallel operation of identical fans has been studied and employed in many plants and facilities. The parallel operation of different fans is far more challenging. Parallel fans may have almost any amount of their operating resistance in common. At one extreme, the fans may have common inlet and discharge plenums. At the other extreme, they may both have considerable individual ductwork of equal or unequal resistance. Fans in parallel should all develop sufficient pressure to overcome the losses in any individual ductwork, as well as the losses in the common portions of the system.

A fan should be selected to match its downstream and upstream systems to provide a proper flow of gas.

When the fans do have individual ducts but are of equal resistance and joined together at equal velocities, they should be selected for the same fan total pressure. If fan velocity pressures are equal, static pressures will be equal. If the two streams join together at unequal velocities, there will be a transfer of momentum from the high-er-velocity stream to the lower velocity stream.


Fans may also be in series but usually at opposite ends of a system. Those in series should handle the same amount of gas by weight measurements, assuming no losses or gains between fans. The combined total pressure will be the sum of the total pressures of individual fans. The volumetric capacities will differ whenever the inlet densities vary from one fan to another. In other words, because of compressibility, the volumetric capacities of the second fan will not equal the volumetric capacities of the first fan. With any fan, the pressure capabilities are also influenced by density.


The case study is for no-load test of a large, critical 2.3 MW fan. This no-load test was part of commissioning to observe and verify if operation and functionality of the fan was suitable for the next stage (load test). The fan was a fixed speed 1,485 rpm turbomachinery which was fed by 11 kV AC power electricity. The rated current was about 140 A and the estimated start-up current was around 750 A.

There were initially different trips on high vibration and some other issues. After resolution of different problems, the test started and operation was sustained. The first problem afterward was a relatively high vibration at the NDE (non-drive end) bearing. The recorded vibration was around 45 microns. This was still below limits published by the manufacturer. However, full investigations were done. Another concern was that this was about two or three times more than other measured vibrations on different bearings.


To start the investigation, the composition of this high vibration (amplitude vs. frequency) was checked. Details of balancing and alignment were also verified. The main part of vibration was related to the operating frequency (1,485 rpm). Therefore, it was assessed as harmless. The manufacturer confirmed that the provisions for rebalancing at site were provided. The key-phasor was provided for this fan and a contractor specialist provided a comprehensive vibration measurement and report that showed no developing problems and no concerning issues.


Relatively high axial vibrations were recorded by handheld vibration measurementdevices on this large, critical fan. Permanent, axial vibration sensors should have been provided for such a large fan. The manufacturer, therefore, was required to install an axial velocity meter.


Another issue was the access to an inspection point of the lubrication oil return line. Sight glass for lubrication oil return piping lines of DE (drive end) side was not easily visible. An additional platform was provided to enable maintenance people to easily check these lines.


The lubrication oil was then found to be dirty. The source of dirt and debris was the system itself due to poor flushing. A complete lubrication oil flushing was done again and verified that all internal parts in contact with the lubrication oil were fully cleaned. ■

Amin Almasi is a Chartered Professional Engineer in Australia and U.K. (M.Sc. and B.Sc in mechanical engineering). He is a senior consultant specializing in rotating equipment, condition monitoring and reliability.

Turbo Tips: Large Fans

For nearly all turbomachinery applications, operators prefer to utilize a direct-driven configuration and avoid gear units. But there are special cases where gear units should be used.

Following an initial evaluation such as a driver speed match analysis, vendors or consultants sometimes inform operators or purchasers that there is no way to avoid the use of a gear unit in a specific turbomachinery train. In high-speed turbomachinery applications, some vendors offer a conventional electric motor with a gear unit. This has been the case particularly for high-speed centrifugal compressors. This is considered by some as an unjustified bias against high-speed, direct-drive electric motors. The use of such motors should be evaluated more closely. The risk and possible operational issues of modern high-speed, direct-drive electric motors are usually less than the problems and issues associated with gear units.

For low-speed turbomachinery trains, the use for a gear unit can easily be avoided. Proper driver types could be selected for machinery options to avoid high-speed drivers for low-speed driven equipment. Steam turbine drivers, though, should usually be limited to direct drive applications. For example, a high-speed steam turbine driver for a low-speed compressor or pump using a gear unit for speed-match is usually a poor decision. Such a complicated configuration can potentially cause many torsional, dynamic and operational issues.

Gas turbine models should be selected to facilitate the direct drive turbomachinery arrangement. Aeroderivative gas turbines usually offer direct-drive trains compared to frame type gas turbines which could only be used in limited speed ranges.

There are tens of thousands of gear units that operate over long periods of time in a wide range of sizes and power-ratings. There are also many gear units for high-gear ratios (above 10), and high-speed applications such as high-speed drivers. Risks, operational issues, and high capital/operational costs associated with gear units have been accepted as fact in many plants. But operators have experienced operational issues, painful installation/commissioning procedures, delays in commission/start-up, numerous unscheduled shutdowns, high cost, and other risks associated with gear units. Because of these issues, there is a growing feeling that gear units should be avoided, if possible.

What is the optimum turbomachinery configuration? It is always best to minimize the number of machinery pieces mounted on a turbomachinery train. That means, eliminating gear units if possible. Doing so, reduces capital, operational, and maintenance costs.

Some gear-unit manufacturers and supporters blame poor application engineering or inadequate maintenance as the reason operators have problems with their gear units. But this does little to lower the relatively high failure rates and issues being experienced with gear units.

Here are a couple of examples from my own experience. They concern integrally geared compressors driven by high-speed steam turbines using intermediate gear units for speed reduction. In such a train, the speed of high-speed steam turbine is first reduced in the intermediate gear reduction. This is used to drive the main input shaft of the compressor. The speed is then increased in the integrally geared train for each compressor casing.

Such a configuration presents inefficient and problematic operation. The speed reduction followed by a speed increase presents considerable losses in the train. It also leads to constant wear and other problems. Such a configuration mandates fixed speed operation of the steam turbine. Instead, this arrangement can be replaced with a more efficient variable-speed steam turbine, direct-driven compressor configuration. This raises the efficiency, reliability, and operational flexibility of the train.

It is true that gear units are needed for many turbomachinery trains. They should be used in cases where other options are not available. But the risks and potential problems associated with gear units should always be considered. Gear units can be eliminated in many turbomachinery trains. Gear units should only be used in turbomachinery trains that other direct drive solutions cannot be employed. 

Should Gear Units Be Used Or Avoided?

Many manufacturers, vendors, engineering firms, and operators have real-ized the importance of lessons learned during different stages of design, devel-opment, and machinery introduction. Too often, those lesson-learned documents have been lists with an overwhelming number of items. Unfortunately, many of the same mistakes are repeated by engineers, vendors, contractors, and operators in subsequent projects. Let’s review a few ways to improve modern turbomachinery. 

The bulk of the design review is done by design experts and engineers. However, there are two groups of people that should be included in each segment of the design review process: the commissioning/start-up team, and the people who will operators operate and maintain the machinery. In many cases a common shortfall is that companies do not select the operations team at the time of machinery selection. It is wise for operating companies to assign operation leads and superin-tendents as soon as possible and have their inputs on all key matters. A similar principle is applicable for the commissioning team. The exchange of data and information between engineering and commissioning is vital. There should be lines of communication estab-lished to help key people communicate each other and transfer vital data, feedback, and experience. 

• High-level requirements by client representatives: It is useful if client representatives establish and explain their high-level requirements and expectations to engineering firms and vendors/manufacturers. Of course, some of the expectations might not be achievable and others might not be reasonable considering technical, budget, or time-schedule constraints. However, it is important to discuss these matters as early as possible. It will be found that while some expectations might not be achievable in the way initially thought, alternatives might be formulated which can achieve those expectations at a reasonable cost and without greatly extending delivery times. It is best to not be in a rush to oppose such expec-tations. It is advisable to spend some time and resources to find a way to achieve them in a more suitable way. 

• Operator feedback: If there was a problem in operation, the best way (and usually the only way) to inform the engineering team (and subsequently vendors/manufacturers) and rectify the problem is via operator feedback. 

• Communication with the commissioning team: The communication between the engineering team (vendors/manufacturers) and commissioning leads is vital. The commis-sioning team need data and documents to plan and prepare for their tasks. Feedback from the commissioning team is also important to consider provisions for smoother and commissioning and start-up practices. Combined with modern monitoring systems, automation provides an opportunity for improved reliability and availability. However, there is such a thing as over-automation. 

Automation is a major trend in plants and facilities. Automated operation is cost effective, efficient, advantageous, productive, and more easy to manage. Combined with modern monitoring systems, automation provides an opportunity for improved reliability and availability. However, there is such a thing as over-automation. 

The basic principle is to leave ordinary, repetitive tasks to control systems and provide operators with tools to do their own tasks better. In an over-automated plant, automation hinders operators from operating efficiently. In some cases, operators cannot intervene and override the automated system when there was a problem. 

A general guideline is to buy new turbomachinery rather than second hand. In many cases, it can cost more to rehabilitate and revamp second-hand turbomachinery that to purchase new equipment. There may also be cracks, corrosion, and erosion that has not been detected. Manufacturer warrantee should be mentioned as a key benefit of new equipment. 

Whatever equipment is used, shop inspection is crucial yet is often overlooked. It should not be left only to third-party inspectors. Such inspectors are beneficial and are nearly always needed. In fact, third-party inspection is often a mandatory requirement. However, inspectors from the oper-ating company, engineer consultants, the main EPC contractor, and others are often needed to inspect machinery at various stages. 

■ Amin Almasi is a Chartered Professional Engineer in Australia and U.K. (M.Sc. and B.Sc. in mechanical engineering). He is a senior consultant specializing in rotating equipment, condition monitoring and reliability

SELECTION, PURCHASE,AND OPERATION OF TURBOMACHINERY

Real-world operational turbomachinery problems can be challenging. Fast decision making is often required. The questions such as “what to do now?” and “how do we deal with this issue?” during live operation can be difficult to answer. It is important to be prepared.

Temperature limits inside a compressor are dependent on compressor design, material selection, and component manufacturing and process gas properties. Mechanical considerations usually differ from compressor to compressor. As a rough indication, temperature limits associated with mechanical design could be about 210°C. However, sealing elements and non-metallic components might limit the gas temperature to around 175°C. But some gases should be kept lower than these temperatures based on process requirements. 

For example, polymerization could be a concern for some gases. If gas temperature passes the limit, polymerization can result in fouling or unbalance masses on impellers which could cause performance deteriorations, long-term vibration, and operational problems. For some process gases, the process limit on temperature could be somewhere between 80°C and 140°C.

A centrifugal compressor with a high flow side-stream was located at the downstream of a complex gas drier system. The side-steam flow was around 3.5 times the main suction flow. During operation at night, a malfunction in the gas drier system raises gas temperature around 40°C compared to normal operation. The compressor experienced a sharp increase in suction temperature. As a result, compressor vibration increased by around 10%, requiring operation team and machinery engineer attention.

A large flow of gas with a relatively low temperature came from the side-stream nozzle and cooled the compressed gas at the mixing section inside the compressor. The suction temperature sharp rise caused an increase of about 9°C at the discharge. Gas temperature increased at the discharge of the first section (before the mixing section) inside the compressor.

Based on measured data, operating parameters were not beyond limits except for the suction temperature. However, there was a concern about high temperatures inside the compressor. An operator recommended tripping the compressor to prevent damage due. The trip was not approved by the machinery engineer and the shift operation leader. Their assessment was that the internal temperature had not passed the limit and any trip could cause a thermal shock in the compressor internals. It would also interrupt production. The operations team acted quickly and corrected the problem in the gas drier system within seven minutes. Suction temperature and vibration levels returned to normal with no sign of damage or malfunction. Additionally, it was confirmed that process gas was not polymerized.

The lesson learned from this event: When the side-stream flow is considerably higher than the main flow, the suction temperature alarm requires careful attention since there could be a possibility that internal temperatures reach a high level and pass the limit while the final discharge temperature may only show a small rise due to the cooling effects of the incoming large volume of side-stream flow.

For an alarm case when all operating parameters are still within allowed limits, the general recommendation is not to shutdown machinery. In the case of increased gas suction temperature for a short time when no parameter passes its limit, the recommendation is to continue operation unless there is a reasonable evidence of the need for a shutdown.

Real-world operational turbomachinery problems can be challenging. Fast decision making is often required. 

Turbo Tips: Tackling real-world challenges 

Before beginning specification, the turbomachinery should be sized. Final sizing will be done by the vendor/manufacturer. But a preliminary sizing is needed for a quality specification: such as specifying a single-stage turbo-compressor when a multistage turbo-compressor is not appropriate. There is often more than one turbomachinery type suited to a defined task. 

For instance, in many modern compression services both integrally-geared turbocompressors and conventional-type turbocompressors could be used. Once the different turbomachinery types are reviewed, various factors sway the decision. The key point is this: It is quite difficult to write a comprehensive multipurpose specification. Therefore, some decisions should be made at an early stage such as the type of turbomachinery. The specification can then be written based on these decisions. That specification outlines what the turbomachinery needs to accomplish and how reliable it should be.

Operational conditions should be divided into a set of normal condition(s) and those other than normal (such as alternative, emergency, etc.). Off-design or unusual conditions such as abnormal gas molecular weight (as might be encountered in a startup), a holding period where part of a plant is not in operation, and similar situations should be accurately included and explained. If anything is missed it could lead to future problems. 

The purchaser generally should know a lot about the plant, the processes involved, and its operation. In particular, the purchaser should be much more aware of unusual conditions, involved risks, and potential upsets. It is unwise to assume that a vendor is completely knowledgeable about the plant or process. 

Reference check any previously successful turbomachinery references in similar facilities. However, each plant or unit has its own set of challenges and specific requirements. Note that turbomachinery manufacturers are specialized in machinery design and manufacturing. Their staff, background, structure, and business are not focused on processes or plant operation.

Off-design or unusual conditions should be included in any specification. 

If the vendor’s engineers have experienced a problem in a process or facility similar to the plant under discussion, they may raise a question. However, this is not often the case. Take fluid temperature runaway potential in hot units. A specification should note some expected maximum temperatures and ask the vendor about the maximum temperature that the turbomachinery can handle. 

Past experience of operation could provide guidance concerning requirements such as the material selection. Stating minimum material requirements can help a vendor in part selection. On the other hand, a specification should not be so rigorous that it seriously limits the vendor. Any specification should specify the minimum material requirements and invite manufacturer input based on their experience.

A large number of unscheduled shutdowns are traced back to vendor design, manufacturing, material selection or component selection such as seal problems, bearing issues, excessive fouling, high degradation, corrosion, and erosion. However, many of these issues could be prevented if the purchaser communicated turbomachinery requirements, expected conditions, and technical specifications properly.

One example could be the selection of austenitic stainless steel. This is a premium material, but cannot be used if chlorides are present due to inter-granular corrosion and subsequent cracking problems. All traces of corrosive, sour or problematic materials should be accurately specified by the purchaser to prevent the selection of wrong material.

There have been many discussions whether to install machinery outdoors or indoors. If in doubt, specify an external location without roof or shelter. But later stages can add shelter or a roof if there is a need to increase overall reliability.

Turbomachinery design parameters, such as turbomachinery speed, should be given as the allowable values after enough research and checking into the most recent technologies and modern turbomachinery installations. These maximum allowable values should be insisted upon to keep bidding compatible.

More time is often spent specifying accessories than is spent on the core turbomachinery, mainly because options are so numerous. Some engineers reported that more than 50% of their discussions and communications with vendors have only been about instrumentations, condition monitoring, electrical systems and control items. All items related to electrical, control and instrumentation should be fully specified early in specifications and clarified at the earliest opportunity. Otherwise, lengthy and challenging discussions are to be expected. Reference: Royce N. Brown, Compressors Selection and Sizing, Third Edition, 2005 (Gulf Publishing, Elsevier).

Click here to sign up for our free eNewsletter.

A step-by-step guide on the turbomachinery specification process, including considerations for unusual conditions and indoor vs. outdoor installation. 

Turbo Tips: How to specify turbomachinery 

Fans have been widely used in many facilities and plants. They have been employed to move gases, been used in different systems to avoid excessive pressure build-up in the equipment or service being served, and to maintain a specified draft over the intended application.

A fan should be selected for the match with its downstream and upstream systems to provide proper flow of gas. This article discusses fans and some practical issues associated with them. The focus is on practical notes and useful guidelines on fans. A case study is also reviewed.

Most experts discourage multiple parallel operation, instead recommending the use of just one operating fan at a time (one operating and another on standby) on each system. This reduces complexity and operational problems as parallel operation has always been complicated and is usually problematic. On the other hand, there have been various reasons for using multiple fans in parallel operation on special systems. In some revamp and renovation projects, two fans may be provided to work in parallel. Capacity control by various fan combinations may be more economical than other control methods in specific cases. Multistage fans may be necessary when pressure requirements exceed the capabilities of a single-stage fan.

Sometimes, two parallel side-by-side fans are close enough to share some items such as lubrication system, shaft, bearings, casing, etc. Double-width, double-inlet fans are essentially two fans in parallel in a common housing. However, the above-mentioned sharing of items or parts should be avoided as it reduces reliability. Having two independent fans is preferred.

Parallel operation of identical fans has been studied and employed in many plants and facilities. The parallel operation of different fans is far more challenging. Parallel fans may have almost any amount of their operating resistance in common. At one extreme, the fans may have common inlet and discharge plenums. At the other extreme, the fans may both have considerable individual ductwork of equal or unequal resistance. Fans in parallel should all develop sufficient pressure to overcome the losses in any individual ductwork, as well as the losses in the common portions of the system.

When fans have individual ducts but are of equal resistance and joined together at equal velocities, they should be selected with the same total pressure. If fan velocity pressures are equal, static pressures will be equal. If the two streams join at unequal velocities, there will be a transfer of momentum from the higher-to the lower-velocity stream.

Fans in series are usually at opposite ends of a system. Fans in series should handle the same amount of gas by weight measurements, assuming no losses or gains between fans. The combined total pressure will be the sum of the pressure of individual fans. The volumetric capacities will differ whenever the inlet densities vary from one fan to another. In other words, because of compressibility, the volumetric capacities of the second fan will not equal the volumetric capacities of the first. With any fan, pressure capabilities are also influenced by density.

The case study is for “no-load test” of a large, critical 2.3 MW fan. This no-load test was part of commission to verify if operation and functionality of the fan was suitable for a load test. This fixed speed 1,485 rpm fan was fed by 11 kV AC power electricity. The rated current was about 140 A and the estimated start-up current was around 750 A.

Initially, there were issues, namely different trips on high vibration. After resolving these problems, the test started and operation was sustained. The first problem afterward was a relatively high vibration at NDE (non-drive end) bearing with a recorded vibration of around 45 microns. While this was still below limits published by the manufacturer, it called for further investigation.

To start the investigation, the compositions of this high vibration (amplitude vs. frequency) were checked. Also details of balancing, alignment, etc., were checked and verified. The main part of vibration related to the operating frequency (1,485 rpm). Therefore, it was assessed as harmless. Also, the manufacturer confirmed that the provisions for re-balancing at site were provided. The key-phasor was provided for this fan and the contractor specialist provided a comprehensive vibration measurement and report including “amplitude – time”, “amplitude – frequency”, graphs, spectrum, etc. The review of the report showed no developing problem and no concerning issue.

Relatively high axial vibrations were recorded by handheld vibration measurement devices. Note that this was a large and critical fan. Yet permanent axial vibration sensors were not provided. The manufacturer was asked to investigate if a kind of permanent axial vibration sensor can be installed on this fan. The manufacturer proceeded to install an axial velocity-meter.

Another major issue was access to an inspection point of the lubrication oil return line. Sight glass for lubrication oil return piping lines on the DE (drive end) side was not accessible. Sight glasses on lubrication oil return piping lines should be accessible by the operator. This major fault was left unchecked from design and installation. It should have been corrected during commissioning. An additional platform was provided for this purpose.

The lubrication oil was found to be dirty. The source was poor flushing. This can cause major difficulties and reliability issues over time. A complete lubrication oil flushing was performed and all internal parts in contact with the lubrication oil were fully flushed and cleaned.

It was observed that the noise was extraordinarily high at the start-up. Noise measurements based on analysis of noise data were performed and no problem was found. A complete noise measurement was planned for the load test.

Turbo Tips: Practical issues with large fans in turbomachinery


Old-fashioned reciprocating compressors have been used in high-pressure services for decades. They have been unreliable, inefficient, heavy, bulky, and wasteful. Modern plants and facilities use vertically-split, dry-gas sealed, shear-ring head centrifugal compressors for high-pressure services (>150 Barg). 

Modern high-pressure turbocompressors should be engineered, manufactured, and operated with a special focus on critical areas such as: rotordynamics, thermodynamics, potential dynamic excitations, rotating-stall, auxiliary systems, fabrication processes, and dry-gas seals. Some examples of high-pressure turbocompressor applications are hydro-cracking, hydrogen-rich process gases (such as those in refineries), methanol plants, gas reinjection (or gas injection), high-pressure transportation pipelines, and some petrochemical processes such as polyethylene processes. High-pressure turbocompressors will continue to play important roles in different industries in the future. Today, competent vendors are capable of manufacturing centrifugal compressors with discharge pressures of more than 900 bar.

A vertically split casing with a heavy bolted head is a classic configuration for medium pressure centrifugal compressors. However, this is not practical for high-pressure turbocompressors because there is not enough space on the casing head for the heavy bolts required. The common configuration for modern high-pressure turbo-compressors is the 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 turbocompressors generally use closed-type three-dimensional impellers. Both the efficiency and the pressure coefficients are higher for three-dimensional impellers compared to conventional ones. Impellers are usually manufactured from low-alloy steels. In-depth material and metallurgical investigations are needed to examine and evaluate the material integrity regarding various failure modes such as sulfide stress cracking (SSC) and others. 

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 high-pressure turbocompressors, the thrust force generated in the impeller assembly is much greater than can be handled in a thrust bearing. Therefore, the thrust forces should be balanced by using double-entry impeller series (a back-to-back arrangement) and balance piston(s). A recommended configuration is a back-to-back impeller arrangement with two balance pistons. The remaining axial (thrust) force is the difference between the forces which are a magnitude larger. A double-acting tilting-pad thrust bearing should be used. 

The shaft-end seals in modern high-pressure turbo-compressors are generally dry-gas seals. Oil seals were used in older machines but are no longer used because of issues such as high power losses, low reliability, oil contamination, and other operational problems. The high pressure presents unique challenges to dry-gas seals. There are only a few dry-gas seal sub-suppliers with good working references for high-pressure services. 

Electric motor drivers have been used in small and medium size turbocompressor trains. A successful configuration is a two-end electric motor arrangement using a dedicated gear unit at each end, to drive two turbocompressor casings (one at each end). Using this configuration, an optimum speed can be achieved for each casing. The compressor bundles can easily be removed from each side. Another configuration is using a two-end, high-speed electric motor with two turbocompressor casings. 

For some high-pressure turbocompressors such as those for gas injection services, gas turbine drivers have traditionally been used. Gas turbine drivers can be employed for compressor trains up to 70 MW (or more).

The first case study involves a high-pressure turbocompressor with a discharge pressure of around 210 Barg and a driver power rating of around 18 MW. It has two casings and is driven by a two-end electric motor. This train employs a variable speed drive to operate at 75%–106% of its normal speed. This offers an excellent operational flexibility, reliability and performance.

Reciprocating compressors used in high-pressure service have been unreliable, inefficient, heavy, bulky, and wasteful.

The second example involves a two-barrel, high-pressure centrifugal compressor. The train consists of a medium-pressure (MP) casing, a high-pressure (HP) casing, 11 MW electric motor driver, and step-up gearbox units. The MP casing (six impellers) receives the gas at around 100 Barg and delivers it at around 225 Barg to the HP casing (five impellers). The HP casing compresses the gas to the final pressure of around 500 Barg.

The third case study involves a complex gas turbine-driven, high-pressure centrifugal compressor for a 580 Barg gas injection service. The train consists of a gas turbine (31 MW), gear unit (the speed increased to around 10,000 rpm), and three compressor casings. The gas pressure increases from 35 Barg to 170 Barg, 360 Barg, and then 580 Barg in three stages (LP, MP, and HP stages). For the MP and HP sections, discharge temperatures are limited to 110°C to reduce the harsh effects of gas, considering the available material technology and long-term reliability required.

Turbo tips: High-pressure turbocompressors

Surge can be a major challenge for turbocompressors. Operation in the surge area will result in an instability, exposing the machine to destructive stresses and forces, high vibration, and even serious damage. 

Surge during shutdown (trip) has been reported for many turbocompressors. This is particularly possible if the machine operates at high head and low flow, immediately before the trip, when the operating point can move toward the surge line and even pass it during coast-down (when the turbo-compressor reduces flowrate). When a turbo-compressor experiences a serious alarm, an emergency shutdown is usually initiated. But an immediate shutdown could result in a surge. In this case, the surge happens shortly after the shutdown (trip) and at a high energy level. This could be a surge at a high head (operating point could pass the surge line at high head). 

In many cases, there are advantages to not removing the driving power from the turbocompressor (tripping) immediately by delaying for a few seconds so the anti-surge valve can be opened and the discharge pressure (head) can be sufficiently reduced. As soon as the trip is intended, the anti-surge is opened, and the compressor shutdown is implemented with a second or two delay. 

Many alarms and malfunctions do not require an instantaneous shutdown. For example, a high bearing temperature and a high vibration (unless it reaches more than 10 times allowable levels). An exception is loss of lubrication oil where it could potentially be worse than a full load surge. A safety study helps determine if such a delay is allowable.

Another critical parameter is the rate of speed reduction (coast-down time) during trip/shutdown. For many turbocompressors, rapid speed reduction can cause surge to be reached sooner and at a higher head condition. This results in a high energy surge event.

Great care should be taken for gas turbine driven compressors. As soon as the fuel supply to the gas turbine driver is cut-off, the power is eliminated to the driven turbo-compressor and the speed drops rapidly. Some installations maintain the fuel flow to the gas turbine driver for up to two seconds, while the anti-surge valve (turbocompressor recycle valve) opens. This delay may generate a safety hazard.

The head-making capability of a turbocompressor is reduced typically by the square of its running speed, while the pressure ratio is imposed by the upstream and downstream piping and facilities system. Therefore, the unit will surge if the anti-surge valve cannot provide fast pressure relief at the discharge system. The deceleration rate as a result of train inertia and energy dissipation is a decisive factor. The rate of pressure relief at discharge not only depends on the reaction time of the anti-surge valve, but also on the volume of gas enclosed in piping and other systems between the compressor and anti-surge valve. 

Speed reduction is fast in aeroderivative gas turbines. An emergency shutdown of a two-shaft or aeroderivative gas turbine-driven compressor can be problematic since train inertia is low and train speed decreases rapidly in a trip. Roughly 20-30% speed reduction can be expected for an aeroderivative-driven machine in the first second after the shutdown. This results in around 50% reduction in head generation capability. An anti-surge valve for such a compressor should be able to reduce the pressure across the turbo-compressor by about half during the same period. The worst-case scenario for an anti-surge system is an emergency shutdown of the gas turbine while the turbocompressor is operating at high pressure and close to surge.

Similar scenarios may be applicable for an electric motor trip. However, the inertia of a typical electric motor-driven train is much higher. A typical rotor assembly for an electric motor driver is relatively heavy and large, and a gear unit is often required. A typical electric motor-driven turbocompressor has three to seven times higher inertia compared to a two-shaft or aeroderivative train.

High-pressure gas trapped in the discharge system plays a major role in surge. Large volumes of pressurized gas need time to depressurize. The volume of pressurized, high-energy gas to be dissipated can be reduced by discharge check valve(s) located upstream of large headers or vessels that store significant amounts of high-pressure gas. Fast-closing check valves are generally specified.

Challenging situations have been reported when the anti-surge valve loop is taken downstream of the aftercooler(s). Usually in these cases, the discharge gas volume in the cooler and piping is too large and the anti-surge valve (recycle valve) cannot avoid a surge. An additional hot gas bypass valve is often required. This is a short recycle loop without any cooler that only operates for a very short time during trip or emergency.

Vent valves on the discharge piping can effectively reduce discharge pressure and stored energy that contributes to the severity of the surge. This is particularly useful in multi-section turbocompressor installations where recycling around the 2nd stage, for example, results in high-pressure gas being added to the 1st stage discharge energy. Venting can allow some gas from the 2nd stage to be removed. Such venting should only be used as the last resort. Most vent valves are small and can be opened rapidly.

Surge can be a major challenge for turbocompressors. Operation in the surge area 
will result in an instability, exposing the machine to destructive stresses and forces, high vibration, and even serious damage. 
