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A step-by-step guide to the revamp of a reciprocating compressor unit to meet increased hydrogen throughput

Revamp of a reciprocating compressor unit

New or expanded process requirements lead to adaptations for compressors. Reciprocating compressors are always ‘tailor made’ for the original operating conditions and capacity. Since processes and/or product specifications have to meet changes in operating conditions, it makes sense to verify existing compressor equipment to see if it is possible to modify or revamp it accordingly.

A reciprocating compressor is designed for a long lifetime. Revamping offers the possibility to use an original investment for changing process requirements and the means to do this is described here.

Current European legislation requires that significant compressor modifications follow the appropriate directives as well as ATEX rules for hazardous installations. The procedures for an ATEX declaration are presented, to demonstrate the investigations required for compressors and accessories. Depending on the kind of modification, the installation of a revamp and the operation of such modified equipment is allowed only with a valid original equipment manufacturer’s declaration according to a machine directive and ATEX.

In 2003 Neuman & Esser (NEA) Group delivered a reciprocating compressor, unit size 2 SZL 320H, to a refinery in Eastern Europe for a desulphurisation process compressing hydrogen from 28 bar suction to 85 bar discharge pressure. The reciprocating compressor is a two-crank, horizontal, two-cylinder, double acting, lubricated service machine (see Figure 1). It has a nominal allowable rod load of 530 kN.  

The compressor is direct driven by an electric motor rigidly coupled, with a nominal drive power of 1700 kW. The original design capacity was 33 000 Nm³/h at a suction temperature of -5°C.

NEA’s scope for the compressor unit included pulsation vessels (suction and discharge side for each stage), interstage cooler, and interstage separator up to the check valve of the last stage.

After five years of compressor operation, demand for hydrogen gas increased due to clean fuels requirements.

An engineering contractor was assigned to prepare detailed specifications based on the new operating conditions and to find the best solution to meet them. The job’s definition was to reach the required capacities even at worst operating conditions.

Verification of compressor feasibility

As-built and pre-check

First of all, the existing reciprocating compressor was recalculated according to the original specification, using a compressor design tool. This verified that the compressor fulfilled the designated process conditions without deviations, to avoid general mechanical or performance problems. The main characteristics of the compressor - dimensions, materials weights and loads - were checked as the basis for further calculations.

Next, a calculation was carried out to check the new operating conditions against the existing compressor design. This showed that the capacity required for new operating conditions could not be reached with the existing compressor. This led to a pre-check in order to see if an adaptation of the compressor for the new operating conditions was possible. The compressor itself would have to be modified, as well as the accessory equipment.

A positive pre-check result meant that there were one or more technically feasible opportunities to match the requirements. There were various possibilities to increase the capacity: by compressor speed, by compressor stroke, and by cylinder bore. 

Preparation of bid

The detailed verification of a compressor revamp is done in the form of an engineering study. In addition to the job specifications, compressor characteristics and details have to be checked based on existing documentation. The acces-sories also need to be checked for the new application. To optimise the revamp solution and to prepare a revamp proposal, different possi-bilities have to be evaluated that will finally result in a tailor made, technically safe and economically responsible solution. 

Feasibility study and revamp possibilities 

Evaluation matrix

To fulfill requirements, the best modification option has to be evaluated. All sorts of technically feasible possibilities are selected and prepared in an evaluation matrix (see Table 1). The matrix shows on one side the different technical solutions for the compressor revamp and on the other side different project features. According to the customer’s ideas, the project features are benchmarked with different quantifiers on a scale 1-10. Position by position, the revamp possibilities are evaluated by multiplying the quantifier of each project feature by the evaluation factor (ok-better-best) of the technical revamp possibility. So in the end, by scoring the most points, the best solution for a revamp is obtained.

According to the scoring, the optimum revamp solution can be selected and used for the following detailed verification.

Thermodynamics and compressor calculations

Detailed verification starts by implementing the specific compressor details regarding loads, dimensions and weights into NEA’s compressor design tool KO³ (Kompressor-Optimierung Version 3). 

According to internal upgrades of different compressors, the year of construction is essential for the layout and the allowable loads. For non-NEA compressors, a similar NEA compressor size is selected. The load limits for the existing compressor have to be adjusted and the main characters for the existing compressor have to be cross-checked and adapted.

The new case data for future operation have to be added to the calculation software. To run the thermodynamic calculation, the gas analysis, suction pressure, suction temperature, discharge pressure and required capacity must be known for each process or case.

To match the required capacity, it is necessary to vary the compressor design figures according to the selected revamp option. To increase the capacity of a reciprocating  compressor, in principle the following possibilities are generally possible, but have to be evaluated according to actual compressor feasibilities.

  • Possibility 1: increase suction pressure - Suction pressure could not be increased due to process conditions.
  • Possibility 2: increase speed - Because of the direct coupled drive, the next asynchronous motor speed was not feasible, since the allowable medium piston speed was exceeded. Also, a speed increase always means a risk of vibration and additional loads on foundations.
  • Possibility 3: increase cylinder diameter - Due to the operating conditions and allowable rod loads, it was not possible to increase the cylinder diameter because the rod loads would be exceeded.
  • Possibility 4: increase compressor stroke - Stroke increase is possible due to the design of the driving mechanism, without modifying the cylinder diameters.
  • Possibility 5: a combination of alternatives - In view of the required capacity, it is enough simply to increase the compressor stroke. A  combination of various alternatives would simply increase costs.
  • Possibility 6: installation of an additional compressor in parallel - There was no additional space to install another compressor.

 

In this case, the conclusion was to install a new compressor crankshaft with enlarged stroke to fulfill the requirements within allowable compressor limits. The piston rods and piston had to be replaced but the cylinders do not need to change, hence the general compressor arrangement can remain.

Verification of compressor valves

Each variation in process conditions or compressor characteristics has an impact on the compressor valves. After detailed calculation of the compressor layout, the valve design needs to be confirmed. Since the valve dynamics have a major influence on compressor performance, a valve check is mandatory. Typically, a slight change in the stroke of the compressor does not influence the valve dynamics drastically. However, it was reported that the compressor suffered from valve problems and lack of delivery rate. The assumption was that oil sticking effects could be responsible for this problem. Therefore, a valve dynamics calculation was done. The valve dynamics tool can also be responsible for oil sticking effects. It was immediately obvious that a late closing of the suction valve due to sticking at the valve guard takes place. This effect leads to a reduction of flow and to higher impact velocities. The effect can be  remedied by stronger spring design in the existing valve.

A new spring design was carried out with the valve manufacturer. In this design, the suction valve starts to close earlier due to the stronger springs. The valve is closed at bottom dead centre and the delivery rate is correct.

In the end the valve size could remain; only the internals had to be adapted according to the new required operating conditions.

Mechanical properties, rod loads, bearing calculation

After the thermodynamics and valves are confirmed, the mechanical properties of the compressor must be verified. For verification, the safety relief valve settings of each stage are an important factor in the compressor’s layout because the set pressures determine the maximum rod forces and static design pressures.

For compressor layout and verification, the mass and gas forces and the combined forces (rod forces) have to be considered. All kinds of forces have to be verified as being inside the allowable loading limits.

Typically, this kind of revamp increases the rod loading of the compressor. Therefore, the standard procedure is to examine the compressor by means of a compressor design program. All compressor parts loaded by the rod load need to be reviewed in terms of possible overloading. The design program KO³ gives a direct impression of the load situation for a given case by generating a bar chart diagram for each rod.

For reciprocating compressors there are sensitive construction groups that have to be checked separately according to operating conditions. The most sensitive compressor construction groups are shown in the form of a bar chart diagram for each rod (see Figures 2 and 3). Figure 2 indicates the different loads for each calculation for Rod 1 and Figure 3 the corresponding loads for Rod 2. 

The maximum allowable loads for NEA compressors are already available for the KO³ calculation. To find a reliable result for non-NEA compressors, the analogies and specific values for dimensions, weights and materials need to be checked and the rod load limits need to be adapted.

In Figures 2 and 3, the yellow bars represent the load utilisation of the individual components in terms of their loading at idle run, meaning the compression chambers are loaded by the suction pressure of the stage. This is more or less the inertia load acting on the components. Typically, this load is not critical for low to mean speed compressors.

The blue bars represent the load utilisation of the individual components in terms of the loading at design condition. The compressor parts are loaded by the gas load due to design suction and discharge pressure and by the inertia load. The red bars represent the load utilisation of individual components in terms of the loading at pressure safety valve (PSV) condition on the discharge side. 

Another aspect concerns the different situations for crosshead-pin bearings at full load and part load operation. It is required that there is always sufficient rod reversal available. Typically, the crosshead bearing is one of the most critical components in a reciprocating compressor. This bearing can fail due to excessive hydrodynamic oil pressure, too small an oil film thickness, or too little rod load reversal. The design program accommodates this by checking these three failure scenarios individually. The calculation showed that the crosshead bearing seems to be slightly overloaded at PSV condition in terms of the minimum oil film thickness for this revamp. The equations behind the load utilisation of the crosshead bearing are derived from a parameter study done with the hydrodynamic bearing calculation tool developed some years ago. The safety margins included in these equations are conservatively chosen. As long as the load utilisation of the crosshead bearing is below 100%, the design is definitely on the safe side. On the other hand, this does not mean that the bearing is overloaded when the load utilisation is above 100%. It just requires more detailed investigation.

For this, an elasto-hydrodynamic (EHD) tool is available to assess the bearing hydrodynamics. This tool is an integral part of the compressor design program. The EHD calculation time for the bearing takes only a few seconds. Assessment of the results is also very simple since one only has to compare the maximum hydrodynamic pressure and the minimum hydrodynamic film thickness with the critical reference values.

A simple representation of the rod loading by a single load would simplify the calculation process since the deformation would be independent of hydrodynamic pressure. Unfortunately, this simple representation is not possible because it would yield such a dramatic deformation that interference between bearing and pin would occur. The calculation process therefore becomes more complex since the resilience of the structure needs to be implicitly introduced in the hydrodynamic solver.

Deformation reduces the hydrodynamic peak pressure compared to a rigid structure since a more equally distributed hydrodynamic pressure takes place along the loaded side of the bearing. Consequently, the minimum oil film thickness also becomes less critical.

According to the calculation, a slight deformation of the bearing takes place. Consequently, the hydrodynamic pressure build-up is concentrated on the central load area between the corresponding oil supply grooves. This is the reason why, even when there is identical compression and tension rod load, the compression side of the cross-head pin bearing is much more critical than the tension side. Unfortunately, the compression load of a reciprocating compressor is typically higher than the tension load.

EHD analysis of the crosshead bearing in terms of hydrodynamic pressure and minimum oil film thickness revealed no critical bearing load for this revamp.

Crankshaft strength

Increasing the stroke of a crankshaft produces higher stress levels which need to be checked in terms of fatigue strength.

The crankshaft load is generally dominated by bending and/or torsion. Their quantification can be carried out by making use of analytical bar models yielding nominal stress levels. Prospective critical locations are the fillets at the crank webs. The evaluation of the stress concentration there is the decisive task in this context. Due to the permanently varying rod load during a crank shaft revolution, the maximum stress in these fillets changes its circumferential position and magnitude all the time.

This effect can be best quantified by utilising finite element analysis (FEA) models. They are used to adjust the local stress situation of the FEA with the analytical bar model by means of stress concentration factors (SCF). Figure 4 shows an example of such a model.

Once a sufficient number of FEA simulations have been performed, their results can be used to identify and adjust analytical approaches which produce approximately the same results as the FEA. That way, the individual crankshaft strength for a given job can be verified most accurately and quickly in the compressor design program without the need for intensive FEA studies. 

The compressor crankshaft in question was modified with an increased stroke. The torsional load at the most critical location had been very low before the revamp and hardly increased afterwards. The bending load, on the other hand, was more dominant in this application and rose approximately proportionally with the increase in stroke. Nevertheless, it remained moderate and left a robust shaft that raised no necessity for any further design adjustment such as using a higher grade material, or increasing the fillet radii or even the pin diameter.

Verification of accessories

Main motor

Each compressor upgrade requires verification of the accessories. The new compressor power must be covered by the existing motor power, or the existing motor must be adapted too. API 618 Rev.5 requires sufficient motor power for each stage against the safety relief valve set pressure, plus a 5% safety margin.

Torsion analysis

For direct coupled compressors, it is mandatory to run a new torsion analysis so that the components of the drive train can be verified to avoid torsion vibrations and compressor damage.

In this revamp case, because of its short length, the compressor crankshaft is barely sensitive to torsional vibrations. This does not significantly change after increasing the stroke. Between the compressor and the induction motor, however, a large flywheel is clamped by two rigid flanges which makes the drive train susceptible to torsional vibrations. The large flywheel had to be installed with the original machinery in order to protect the electric power supply grid against current fluctuations which are inherent in every motor driving a reciprocating compressor. So the system is basically a two-mass oscillator capable of torsion within the motor shaft section between flywheel and rotor core. The natural frequency of this is not affected by a changed crank shaft stroke. It is only the torsional excitation that rises to some extent. Therefore the torsional vibration simulation had to be rechecked for the revamp and everything was found to be within acceptable levels.

Gas coolers, vessels, piping

If there is an increase in capacity and power, the coolers have to be verified for the new operating conditions. These detail checks need to be done by the original equipment manufacturer of the heat exchangers. Based on utilities and design limits, the equipment is recalculated for the new operating conditions.

For a new flow or new operating conditions, the sizing and ratings of piping and vessels need to be confirmed. To avoid excessively high pulsation, a damper check is performed based on the existing vessel design and according to the limits in API 618. The damper check can be run with a KO³ calculation and considers the new compressor layout’s overall processes and operating cases. The result provides first information as to whether the existing vessels are still sufficient or too small.

The damper check by KO³ software only gives preliminary results for expected pulsations after the revamp. The software gives a proposal for pulsation damper design.

Finally, for detailed verification and for providing a warranty regarding vibrations according to API 618 and NEA workshop stand-ards, it is necessary to run a pulsation study and mechanical response study.

Pulsation study

A pulsation and mechanical response study was also carried out for the original compressor design in this revamp case. The modifications to compressor design and thermodynamic operating conditions were communicated to the supplier to reinvestigate and validate the system. For the pulsation study, the cases to be investigated are single and parallel operation, the different operating conditions at normal situation (100%) capacity and also for the capacity control range between 100-15% by stepless suction valve unloading. According to experience with synchronised suction valve unloading, the critical point for pulsation volume design is at approximately 70% capacity. If the study requires it, the modifications to the system to satisfy acceptable operation, by order of preference, are:

  • Modification of restriction orifices
  • Modification of existing piping supports
  • Modification of pulsation dampers

In this case, the calculation showed excessive pulsation levels as well as unbalanced forces on the suction side first stage for the existing vessel volume. Therefore a new acoustic filter had to be considered with increased volume and dimensions. Also, a steel construction to support the new vessel and new orifices for the line side had to be considered.

Machine directive/ATEX

According to machine directive 2006/42/EC and directive 94/9/EC (ATEX) it is essential, in the event of a substantial modification, to provide a manufacturer’s declaration for the modified equipment. The operating company has to ensure conformity with the machine directive and ATEX for the whole operation. This is also mandatory for used machines which are substantially modified in a revamp. ATEX does not apply to repairs without new features or other modifications, or for spare parts intended to replace defective or worn out parts.

In this revamp case, there is a capacity increase within the original operating conditions through the installation of a new crankshaft with an enlarged stroke. The crankcase and other internals of the driving mechanism remain. Also, the first and second stage cylinders of the compressor remain. To achieve the enlarged stroke, a new piston and piston rods were designed. Other compressor characteristics such as speed and cylinder bore remain identical to the original design.

To run the compressor under high suction pressure and at low suction temperature, the driver had to be renewed.

The oil system, the control system and other equipment was already certified according to ATEX when the compressor unit was installed and did not need to be modified. The distributed control system (DCS) and the compressor instrumentation for safety control could remain; only the set points in the instrument list had to be adapted.

Does a substantial modification need to be considered? The following questions have to be carefully checked:

  • Are there changes in function or operating conditions for the compressor unit?
  • Are all hazards covered by risk analysis?
  • Or do new risks or new hazardous conditions exist after the revamp?
  • What about the effects or rate of incidents arising from existing hazards?

Figure 5 shows an evaluation for the original equipment manufacturer and operating company to check whether a substantial modification is occurring.

The oil system, the control system and other equipment was already certified according to ATEX when the compressor unit was installed and did not need to be modified. The distributed control system (DCS) and the compressor instrumentation for safety control could remain; only the set points in the instrument list had to be adapted.

Does a substantial modification need to be considered? The following questions have to be carefully checked:

  • Are there changes in function or operating conditions for the compressor unit?
  • Are all hazards covered by risk analysis?
  • Or do new risks or new hazardous conditions exist after the revamp?
  • What about the effects or rate of incidents arising from existing hazards?

Figure 5 shows an evaluation for the original equipment manufacturer and operating company to check whether a substantial modification is occurring.

Evaluation by risk analysis

NEA Group has generated a spark hazard analysis and risk assessment for reciprocating compressor units. If there is substantial modification in a revamp, the assessment is performed and actions are indicated. When there is full documentation for compressor data, materials and design available, it is possible to prepare a declaration according to the machine directive and ATEX. The reciprocating compressors can be confirmed for CE Ex II 2G T3.

Results of analysis

For this revamp there are:

  • No new hazardous events compared with the previous situation
  • None of the existing hazardous events increased; the new stroke was recalculated and the loads checked by the original equipment manufacturer. This means that existing safety protection would have been sufficient even after the revamp 
  • Besides compressor layout, all hazardous events were covered by the existing risk analysis:
    • incident rates and effects of pressure and temperature shift
    • valve failures
    • pressure packings
    • rod-reversals
    • crosshead and bearings
    • drive components
    • oil pump/oil system
    • compressor function and safety system is identical as before the revamp
    • boundaries and interaction with bordering systems
    • compressor remains within the same operational environment; classification and configuration of hazardous areas do not change.

Risk evaluation is mandatory for each revamp project and must be performed during the revamp project phase. The results of the evaluation must be documented, for proof. 

For revamps and modernisations, special management of product safety is necessary. The original equipment manufacturer or an expert authority must state that the revamp measures are a safe solution. How the revamp is selected and designed, as well as the associated risk analysis, must be documented in detail.

Conclusion

Revamp and modernisation of reciprocating compressors requires special handling. Because a lot of reciprocating compressors have been running for decades, and may even run for decades more, a revamp is a good opportunity to ensure that they fit current technical specifications and process conditions. With the right technical support by a compressor OEM, a reciprocating compressor can be prepared for long term operation, meeting the demands of the operator company.

A revamp proposal will demonstrate a technically safe and economically reasonable solution using a major component of existing resources. Beside the compressor and the compressor equipment itself, infrastructure such as foundations, compressor housing, cable routing, pipe racks and so on may be further used.

A major revamp does not automatically require substantial modification, but extensive modification can arise in smaller revamps if there is a change in risk assessments. Only an OEM or an expert authority with sophisticated engineering knowledge of reciprocating compressors can provide the necessary manufacturer’s declaration or documentation so that the operating company can receive  permission to put the revamped unit into operation.

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