Experiences with rotordynamic stability on High Pressure ... · Experiences with rotordynamic...

XIX International Gas Convention – Caracas, May 24-26, 2010

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Experiences with rotordynamic stability on High Pressure Centrifugal Compressors (equipped with Hole Pattern): Design, Calculation and Testing

Y. Bidaut (Manager Mechanical Development) [email protected]

U. Baumann (Head of Calculation and Development) [email protected]

MAN TURBO AG Schweiz Hardstrasse 319

CH- 8005 Zürich, Switzerland

1. Introduction

In recent years, applications such as enhanced oil recovery (EOR) have created

an increased demand for high pressure injection compressors, very often in

combination with wet and sour gases. These compressors represent a high

challenge with respect to the stability. Worldwide the majority of the compressors

are operating at discharge pressures below 300 bars (average densities below

approx. 200 kg/m3). Only a small number of compressor run at pressures above

400 bars and average densities above 250 kg/m3. To ensure the stability of

modern medium and high pressure compressors devices such as swirl brakes and

damper seals are used. The stability behaviour of the compressor equipped with

such damper seal is dominated by the forces produced by this device. Therefore in

the last decade hole pattern damper seals have become a standard design feature

in High Pressure Centrifugal Compressor.

2. High Pressure Compressor design

Fig. 1 shows a 3D drawing of a typical High Pressure compressor. For

aerodynamic stability and enhanced pressure ratio all diffusers are of a bladed

design. The stages are equipped with thrust brakes. These devices take out the

circumferential speed component of the gas in the shroud side room and

consequently change the pressure within this cavity. The effect leads to a smaller

stage thrust and to a smaller thrust variation for different operating and seal

conditions. As a second effect thrust brakes reduce the inlet swirl to the shroud


.

labyrinths and are therefore considered a stability increasing feature. The stages

are equipped with stationary see-through labyrinth seals. For the balance piston a

hole pattern design is used. The sleeve consists of a single circular piece with a

convergent tapered bore.

Fig.1: 3D-Model of a high pressure compressor with drawing of an impeller with

thrust brake and a photo of the hole pattern seal

Earlier high pressure turbocompressors were equipped with normal labyrinth seals

at the balance piston. Compared with hole patterns, labyrinth seals provide much

lower levels of forces and stiffness. Therefore the forces produced by labyrinths

are much less sensitive to clearance changes and diverging gaps. For the earliest

hole pattern applications the seal design was the same as for the labyrinth seals

(balance piston stator sleeve mounted on the inner casing). However this design

provided a lot of flexibility that resulted in large deflections that are detrimental to

the stability of the seal gap geometry: A severe tilting of the entire inner casing end

diaphragm occurred, leading to a very large taper change for this configuration.

Therefore to overcome the various disadvantages of this configuration, improved

casing designs were developed (unsegmented stator sleeve mounted directly onto

the cover).


.

3. Calculations

3.1. Lateral Analysis

The accurate determination of the compressor’s stability is necessary for a correct

layout. This implies a profound knowledge of the parameters that influence the

rotordynamics. The most dominant component affecting the rotor stability is the

hole pattern seal. The comparison between the first bending modes for the

unloaded and the loaded compressor shows the influence of the damper seal

acting on the compressor like an additional bearing (Fig. 2). An adaptation of the

hole pattern parameters is performed in order to reach the best accuracy.

Fig.2: Mode shape of the lowest bending mode for the unloaded and loaded

compressor

3.2. FE Analysis

In order to quantify the rigidity of the casing design a finite element analysis (FEA)

is performed. The most important parameter to be considered is the deformation of

the casing at operating conditions. Special attention must be paid to the clearance

of the hole pattern seal that must be kept convergent at all operating conditions. It

is desired to have a convergent (inlet larger than exit) conical clearance for the

hole pattern seals to prevent the strong negative direct stiffness that occurs in the

hole pattern seal with diverging clearance. Due to large differential pressures

across the balance piston seal the manufactured clearances change in operation.

Fig. 3 shows the total deformation of the casing parts due to pressure only. It turns

out that increasing pressures lead to a growing tendency towards divergence at

the hole pattern seal. Similarly the thermal influence is analysed taking into

account the thermal distribution in the casing. Furthermore the reduction of the

clearance due to the centrifugal forces of the rotor is considered. Hence the slope


.

of deformation with respect to pressure and temperature difference is used later to

calibrate the predicted hole pattern rotordynamic coefficients.

Fig.3: Deformation (entire casing: total, balance piston detail: radial) of the casing

due to pressure loading

3.3. Stability Analysis

To predict the frequency dependent stiffness and damping coefficients of the hole

pattern seal the code developed by Kleynhans and Childs [1] is used. The

coefficients of all labyrinth seals are calculated using a combination of bulk-flow

and CFD software. Thereafter, the stability (natural frequencies and damping) is

calculated for each operating point. Fig.4 shows the lowest logarithmic decrement

value for any mode with a whirl frequency below 300 Hz as function of assumed

taper amounts. The figure clearly shows that the compressor operates in a stable

regime. The considerable margin between the “seal taper stability threshold” and

the actual taper values demonstrates the robustness of the design.

Fig. 4: Sensitivity to hole pattern seal taper

-400

-300

-200

-100

0

100

200

-150 -100 -50 0 50 100 150 200

Seal Taper (μm)

Log.

Dec

. (%

)

Divergent Convergent

cold, design

in operation

stableunstable

stab

ility

thre

shol

d

Separation Margin to stability threshold


.

4. Measurements

4.1. Generally

MAN Turbo never encountered any unstable High-Pressure Compressor.

Nevertheless Full-Load, Full-Density (FLFD) tests are performed on some

compressors for following purposes:

- Verification of the prediction

- More accurate design Rules

- Improvement of efficiency (only as stable as necessary)

- Enhanced customer confidence and satisfaction.

The mechanical and thermodynamic measurements are performed in the same

test setup in accordance with the API 617 (2003) and ASME PTC 10 (1997) Class

II specifications. Prior to these tests the FLFD test is performed on the

compressor. During the high-pressure test the vibrations, damping ratio, thrust,

and the leakage at the balance piston seal are measured. Furthermore, a

frequency analysis is carried out to verify the absence of sub-synchronous shaft

vibration. In order to reach a similar gas discharge density as on site pure nitrogen

(N2) or a mixture of nitrogen and carbon dioxide (CO2) is supplied to the loop.

The main flow parameters influencing the seal behaviour are the pressure, the

density and the leakage mass flow (dependent from the pressure and the density).

Hence with a gas especially chosen for a mechanical full-load, full-density test

these parameters can be adjusted very well.

Fig. 5 shows a reference chart containing the highest pressure applications built to

date on the API Stability diagram as well as the MAN Turbo compressors

equipped with hole pattern seals.

On the following the paper will show one example of a FLFD tested compressor

(Ref. 1) and one example of a FLFD and Stability measurement of a 655 bar

compressor (Ref. 2).


.

Fig. 5: API 617 Level I Diagram with High Pressure Compressor References

4.2. FLFD- Test of a 340 bar - Compressor (Ref. 1)

This year MAN Turbo delivered 2 HP-trains consisting each of 2 single

compressors (HP1 and HP2) for an offshore application near the coast of Angola.

In November 2009 FLFD tests were performed at each compressor. Fig. 6 shows

the original train arrangement of the skid tested in the test bed facility in Zürich.

Fig.6 : Compressor Train Arrangement (Ref.1)

The comparison of the density and the pressure along the stages of the HP2-

compressor between the operating on site end the testing (Fig.7) shows that the

FLFD test fully reflects the specified conditions. The deviations of the test pressure

and density from the specified values are in the same range than those allowed for

class 1 tests.

HP2 (Ref.1) HP1

1.0

1.5

2.0

2.5

3.0

0 100 200 300 400 500Average Gas density (kg/m3)

Crit

ical

Spe

ed R

atio

CSR

Level I

Level II

FLFD tested

MAN Turbo Compressors with Hole Pattern

FLFD tested & Damping Measurements

Ref.2

Ref.1


.

Fig.7: Pressure and density at FLFD-test and for specification (HP2-compressor)

The waterfall plot, represented in Fig. 8, shows no subsynchronous vibrations

during this test, demonstrating the stability of the compressor.

Fig. 8: Waterfall Plot during FLFD-Test

4.3. FLFD and Damping Measurements of a 655 bar - Compressor (Ref. 2)

For an application in Oman MAN Turbo recently supplied an injection compressor.

Before shipment to site, the compressor was subjected to comprehensive testing

in the MAN Turbo facilities in Zurich. The stability tests were performed up to a

maximum discharge pressure of 655 bar. The machine has a pressure rating of

700 bar. The layout accounts for two different gear sets, a main 100 percent gear

ratio for all specified operating conditions and a spare 105 percent gear set as a

provision for lower molecular weights or higher required pressure ratios. Tab.1

shows the main operating conditions of the compressor for site operation as well

as for the FLFD-shop tests.

HP2 – DE (suction)

HP2 – NDE (discharge)

100

150

200

250

300

1 2 3 4 5 6

Stage N° and Volute

Test (N2= 100%)

Guarantee (Natural Gas)

Density (kg/m3)

100

150

200

250

300

350

400

1 2 3 4 5 6

Stage N° and Volute

Pressure (bara)


.

Tab. 1: Compressor operating and test conditions

The train tested in the original equipment manufacturer’s test bed consists of a

synchronous motor driving, via a speed increasing gear, the HP compressor (Fig. 9).

Fig.9: Compressor Train Arrangement

To measure the frequency and damping ratio of the lowest bending mode at full

speed, the first natural frequency was excited asynchronously with a magnetic

bearing. Fig.10 shows a solid model assembly of the end of the rotor with the

magnetic bearing exciter installed. The shaft response was measured while

running the magnetic force excitation frequency through the natural frequency of

the rotor.

A BSuction pressure bara 233 288 300Discharge pressure bara 560 560 648Average Gas density kg/m3 366 301 384Discharge Gas density kg/m3 404 327 426Rotor speed (100%) rpm 12068 12068 12068

CO2+CnHm CnHm+CO2 CO2+N2

g/mol 28.2 23.2 30.7

Test

Gas

Operating conditionsFeature Unit

Synchr. Motor

P = 11‘500 kWn = 1‘500 rpm

Injection Compressor

n100% = 12‘068 rpmn105% = 12‘643 rpm

Gear box

Flexible Coupling

Solid Coupling

Synchr. Motor

P = 11‘500 kWn = 1‘500 rpm

Injection Compressor

n100% = 12‘068 rpmn105% = 12‘643 rpm

Gear box

Flexible Coupling

Solid Coupling


.

Fig.10: Model of free end of compressor with Magnetic Bearing Exciter

Procedure

At the maximum rotating speed (12’643 rpm), the damping ratio and the thrust

were measured along the performance curve from near surge down to near choke

with constant suction pressure of 300 bar. The suction pressure was then reduced

to 250 bar, and the equivalent measurements were performed for this performance

curve. Afterwards the operating point was adjusted near the rated point, and the

suction pressure was reduced in steps of 50 bar down to 50 bar. Additional

measurements were performed with a suction pressure of 25 bar and 10 bar. At a

rotating speed of 12’068 rpm the measurements were performed along the

performance curve for suction pressures of 300 bar and 250 bar (Fig. 11). The

damping ratio measurements were carried out by applying a harmonic force with a

sweep over a frequency range from 50 Hz to 300 Hz.

Fig. 11: Stability measurements at these operating points

Magnetic Bearing

Additional Extension

Shaft

1.4

1.6

1.8

2.0

2.2

2.4

0.7 0.8 0.9 1.0 1.1 1.2 1.3 1.4

Suction Volume Flow (normalized), Vs / Vs,rated

Pres

sure

Rat

io P

I

105% Speed / Gas : N2 (100%)

100% Speed / Gas : N2 (83%) + CO2 (17%) ps= 310 bara

ps= 250 bara

ps= 250 bara

ps= 300 bara

ps= 10 bara

ps= 200 bara

ps : Suction Pressure

. .


.

Measurements

Fig. 12 shows a waterfall plot of the loaded compressor while exciting the forward

precession mode. The extracted orbit at the critical speeds indicates the

precessional direction. The recorded averaged time series of the measured shaft

vibrations were transformed into the electromagnet axes and then FFT-

transformed and thereby, resonance peaks –if existent- could be recognized. The

damping ratio was evaluated by the circle fit method.

Fig. 12: Waterfall plot while exciting the forward precession mode

The results of the stability measurements (Fig. 13) show that the compressor is

very stable throughout the entire performance map. The most remarkable result on

the stability curves is the large rise of the natural frequency at low pressure

(increase from 100 Hz [unloaded] up to 140 Hz [discharge pressure of only 25

bara]). The natural frequency then increases steadily with increasing pressure up

to 180 Hz at 300 bar. Beyond this pressure the natural frequency did not increase.

The same observation can be made for the damping ratio: The damping ratio

already increases at very low pressures reaching a constant value of about 120 %

log dec at 300 bar. The stability behaviour of the compressor is more or less

independent of the suction pressure at constant discharge pressure. As expected

the damping ratio decreases with increased rotational speed. The comparison of

these test results with the measurements performed on high-pressure

x-direction

y-di

rect

ion

ωt = 0 ωt = π/2

Rotor


.

compressors equipped with a labyrinth balance piston [2] reveals the major

influence of the hole pattern seal on the natural frequency. In contrast to the

labyrinth piston, the damper seal, which increases the natural frequency with

increasing discharge pressure, ensures a constant high stability of the

compressor.

Fig. 13: Measured and Predicted Natural Frequency and Damping versus Discharge

Pressure

From the comparison between the measurements and the prediction it can be

concluded that the seal forces in radial direction are calculated with a high level of

reliability. At low discharge pressures the frequency is higher than expected,

whereas at high discharge pressures the calculated natural frequency is

approximately 10 Hz higher than the measured value. Regarding the damping

ratio, the predictions give a conservative result.

6. Conclusions

The full-Load, full-Density and the stability tests presented above demonstrate the

very satisfactory damping of the compressors and hence the robustness of the

design.

The results presented in this paper show that high-pressure compressors are well

damped (above 100 percent logarithmic decrement) over the complete tested

discharge pressure range up to 655 bar (corresponding to 384 kg/m3).

80

100

120

140

160

180

200

220

0 100 200 300 400 500 600 700Discharge Pressure (bara)

Calculation - 105% SpeedCalculation - 100% SpeedMeasurement - 105% SpeedMeasurement - 100% Speed

Natural Frequency (Hz)

0

40

80

120

160

200

0 100 200 300 400 500 600 700Discharge Pressure (bara)

Log Dec. (%)


.

The comparison of predictions versus measurements shows that the damping is

well predicted even though somewhat underestimated for discharge pressures

above 300 bar. The first bending frequency is well predicted for discharge

pressures above 100 bar.

The margin to the stability threshold is high enough to optimize the compressor

with respect to the leakage and performance without impact to the rotordynamic.

Due to the experience gained by the successful testing and analyses of this

compressor, MAN Turbo plans to develop compressors beyond 800 bars.

7. References

[1] Kleynhans, G. F. and Childs, D. W., The acoustic Influence of Cell Depth on the

Rotordynamic Characteristics of Smooth-Rotor/Honeycomb-Stator Annular Gas

Seals, ASME Journal of Engineering for Gas Turbines and Power, 1997

[2] Baumann, U., “Rotordynamic Stability Tests on High-Pressure Radial

Compressors”, Proceedings of the twenty-eighth Turbomachinery Symposium

College Station, Texas, 1999

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