Mantosh Bhattacharya

Centrifugal compressors have lubrication systems designed as per API 614 consisting of an oil reservoir, pumps, filter, cooler and oil lines. The programmable logic controller (PLC) starts the auxiliary AC power-driven lubricating oil pump.Pressure at the oil header creates starts the drive.If the header pressure is low, the standby oil pump starts and maintains oil pressure during operation. Oil is pumped to the lubricated parts (radial and thrust bearings). If a speed increaser gear is present, they are lubricated with the same oil as the bearings (Figure 1).

 Figure 1: Generic lube oil system diagram for a centrifugal compressor[/caption]

Some end users require two pumps that have different types of drive or power sources. When both are driven electrically, they have to be connected to separate supply feeders in order to maintain constant and uninterrupted operation should a feeder fail.However, it is not required for oil coolers to have such redundancy.

In some compressors with step-up gearboxes, the main oil pump is required to be driven mechanically from the gearbox, and the auxiliary pump operates during the start-up and run-down phases of the compressor train. As API 614 calls for the oil reservoir to be away from the skid with a slope requirement, shaft-driven mechanical pump performance may be hampered by net positive suction head-related (NPSH) air bubble issues (Figure 2).

 Figure 2: API 614 compliant lubricating oil system[/caption]

Sometime a screw pump connected to a gearbox can cause rotor-dynamic problems due to high overhung weight. Shaft-driven gear pumps are mostly used to maintain oil supply during coast down in the case of an electrical power outage as compressors need lubrication until their rotating shafts come to a complete stop.

A conventional approach is to maintain lubrication for a fixed period of time following a shutdown request. Based upon an estimate of the time for a rotating shaft to stop,this approach, does not always ensure that the shaft receives enough lubrication. Therefore, a rotating shaft and its corresponding bearings may be physically damaged if it rotates after that fixed period. To avoid possible issues,therefore, a well sized run-down tank is a better alternative.

Another big lubrication challenge is avoiding bearing damage during reverse rotation. This can occur during coast down, gas blow down, or short circuiting of high-pressure gas into a standby compressor.Most of the time,radial bearing pads have an offset pivot design while thrustbearings have a tapered land-thrust design that has a significantly lower load-carrying capacity in the reverse direction. The compressor designer can specify center-pivoted pads in bearings and bi-directional seals. But regardless of the design, rotor train inertia must be higher than gas inertia.This cannot be guaranteed for a high-pressure compressor if it has a low rotor weight.In that case, due diligence is required to overdesign the overhead tank OHT to meet the duration of reverse rotation,and to create a short circuit between the suction and discharge lines with a class IV sealing valve (open when the compressor is not in operation) or a fast-acting recycle valve.

If the compressor is a single operating unit with no standby, it is advisable to start the lube oil pump first before beginning blow down of the compressor loop. The duration of operation of the lube oil pump is dictated by API 521 as regards the blow down rate of loop. A non-slam check valve at the discharge side is usually installed although a clause of API 521 states that a single check valve cannot be regarded as a reliable layer of protection against reverse flow.

Another option is to provide a reverse rotation detector which can give a command to start a DC pump or the installed auxiliary AC motor driven pump.The reverse rotation detector can be set to arm a digital tachometer connected to a microprocessor or PLC.The tachometer should be suitable for low RPM detection. The RPM signal from the tachometer is transmitted to the PLC which can the start the auxiliary or DC pump.All of the above points must be taken into account while placing an order for a turbo-compressor package.

Mantosh Bhattacharya is a rotating equipment specialist based in the United Arab Emirates.

Articles

Configuring lubrication systems

In a recent site acceptance test, it was found that the shape and position of the centrifugal compressor curves varied from the design performance curve. These compressor curves were taken from the compressor as tested curve and contractual data sheet, and then used to calculate polytropic head and actual inlet flow for 12 test data points, at over 4 different compressor speeds. These values were further verified using software (HYSYS) calculation.

The performance points were plotted on the design performance curves. In addition, the fan laws were used to calculate additional ‘design’ speed curves, at compressor speeds equal to those used to gather the performance data. This allowed a direct comparison between the measured performance points and the calculated design speed curves. The plot is shown below -

Blue lines – Design constant speed curves from built performance curves at 70%, 79%, 89%, 92%, 100% and 105% speeds respectively

Black lines – Design Surge Control Line (SCL) and Surge Limit Line (SLL) from built performance curves

Red solid lines – Calculated constant speed lines at 72%, 80%, 90% and 95% speeds

Red Crosses, connected by Red dashed line – Performance data points from on-site performance testing at 72%, 80%, 90% and 95% speeds

Circled Crosses – Performance points with a Compressor controller DEV of almost zero, expected to be on the SCL.

The performance data at site correlated well with the design compressor curves. This was indicated by a DEV value on the compressor controller of 0.05 (i.e. almost 0) for two of the points taken (circled). This raised doubts regarding the correct set-up and positioning of the Compressor controller SCL.

Compressor controller surge control line set-up

The set-up of the compressor controller SCL was investigated, particularly the calculation of flowrate from the V-Cone flowmeter device pressure drop (deltaP) reading. It was seen that for a given deltaP, the calculation of mass flowrate by the compressor controller did not correlate with the calculated mass flowrate by process engineering department.

For example, flow measuring device (FMD) delta P of X bar at X bara suction pressure:

Compressor controller equated this to Y kg/h

Process engineering department equated this to Z kg/h

It was found from the compressor controller manual that the system calculates flowrate from the pressure drop across the flowmeter device, and it is dependent on a constant A, as per the equation:

ΔP = pressure drop across the flowmeter device

FMD A constant is calculated using a number of factors and coefficients specific to the flowmeter device sizing, as well as the service conditions:

This equation enables the conversion of pressure drop across the device into flowrate in the compressor controller software. However, the values used in this formula as stated in the compressor controller manual, differed from the values stated on the flow measurement device data sheet. It was, therefore, suspected that compressor controller could have been supplied with incorrect flowmeter device sizing data.

To correct this error, an inaccurate constant ‘A’ was being used within the compressor controller software, and would subsequently lead to the incorrect calculation of flowrate by the compressor controller. This led to the compressor controller SCL being set up significantly to the right of the design SCL.

The compressor manufacturer and the compressor controller software supplier, along  with the site process engineering department support, worked out the embedded errors in the calculations and resolved all discrepancies in the calculation of flowrate.

First, the nameplate of the flowmeter device in the field was reviewed to ascertain the correct device sizing. This data was used to calculate and check the compressor controller ‘A’ constant and ensure that the correct ‘A’ constant is used within the compressor controller software. The software will need to be updated in the field as required.

The compressor controller testing was then repeated to correctly set the Surge Control Line and shaping of compressor curve.