development and veriﬁcation of 3000rpm 48inch integral

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Development and Verification of 3 000 Rpm 48 Inch Integral

Shroud Blade for Steam Turbine∗

Yasutomo KANEKO∗∗, Kazushi MORI∗∗ and Hiroharu OHYAMA∗∗∗

The 3 000 rpm 48 inch blade for steam turbine was developed as one of the new stan-dard series of LP end blades. The new LP end blades are characterized by the ISB (IntegralShroud Blade) structure. In the ISB structure, blades are continuously coupled by blade un-twist due to centrifugal force when the blades rotate at high speed. Therefore, the numberof the resonant vibration modes can be reduced by virtue of the vibration characteristics ofthe circumferentially continuous blades, and the resonant stress can be decreased due to theadditional friction damping generated at shrouds and stubs. In order to develop the 3 000 rpm48 inch blade, the latest analysis methods to predict the vibration characteristics of the ISBstructure were applied, after confirming their validity to the blade design. Moreover, the ver-ification tests such as rotational vibration tests and model turbine tests were carried out inthe shop to confirm the reliability of the developed blade. As the final verification test, thefield test of the actual steam turbine was carried out in the site during the trial operation, andthe vibration stress of the 3 000 rpm 48 inch blade was measured by use of telemetry sys-tem. In the field test, the vibratory stress of the blade was measured under various operatingconditions for more than one month. This paper first presents the up-to-date design technol-ogy applied to the design of the 3 000 rpm 48 inch blade. In the second place, the results ofthe various verification tests carried out in the shop are presented as well as their procedure.Lastly, the results of the final verification tests of 3 000 rpm 48 inch blade carried out in thesite are presented.

Key Words: Steam Turbine, Blade Vibration, Verification Test

1. Introduction

New standard series of LP end blades of steam turbinehave been developed to improve the thermal efficiency andthe reliability. The new LP end blades are characterizedby the ISB (Integral Blade Structure)(1), (2). More than 15LP end blades including the 3 600 rpm 45 inch titaniumblade and 1 500/1 800 rpm 54 inch blade have applied theISB structure, and shown the superior experience. The3 000 rpm 48 inch blade for steam turbine was developedas one of the new standard series of LP end blades to meetthe demand of the larger capacity and the lower cost. Indeveloping the 3 000 rpm 48 inch blade, the latest analy-sis methods to predict the vibration characteristics of the

∗ Received 3rd October, 2005 (No. 05-4193)∗∗ Takasago R&D Center, Mitsubishi Heavy Industries, Ltd.,

Takasago, Hyogo 676–8686, Japan.E-mail: kazushi [email protected]

∗∗∗ Takasago Machinery Works, Mitsubishi Heavy Industries,Ltd., Takasago, Hyogo 676–8686, Japan

ISB structure were applied, after confirming their validityto the blade design(3), (4). Moreover, the verification testssuch as rotational vibration tests and model turbine testswere carried out in the shop to confirm the reliability ofthe developed blade. As the final verification test, the fieldtest of the actual steam turbine was carried out in the siteduring the trial operation, and the vibration stress of the3 000 rpm 48 inch blade was measured by use of telemetrysystem. In the filed test, the vibratory stress of the bladewas measured under various operating conditions for morethan one month.

This paper first presents the up-to-date design tech-nology applied to the design of the 3 000 rpm 48 inchblade. In the second place, the results of the various verifi-cation tests carried out in the shop are presented as well astheir procedure. Lastly, the results of the final verificationtests of 3 000 rpm 48 inch blade carried out in the site arepresented.

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2. Design of the 3 000 Rpm 48 Inch Blade

The length of the low pressure end blade is the mostimportant factor in designing the high efficiency and lowcost steam turbine, because it determines the turbine out-put, the performance, the required number of casing, andso on. In design of the 3 000 rpm 48 inch blade, the bossratio, which is the ratio of the blade base diameter and theblade tip diameter, was reduced to 0.4 from 0.5, whichhas been used in the conventional blade design, in or-der to increase the blade length and the exhaust annulusarea and decrease the dimension of the casing. As a re-sult, without changing the blade material from the con-ventional 17-4 PH steel (stainless steel), the 3 000 rpm48 inch blade achieved the 1.2 times longer blade and 1.3times larger exhaust annulus area, comparing to the con-ventional 3 000 rpm 40 inch steel blade. Therefore, apply-ing the 3 000 rpm 48 inch blade to 50 Hz-600 MW classsteam turbine, the number of the casings of the low pres-sure turbine can be reduced to one (2 flow type) from two(4 flow type). As for the steam turbine of the tandem com-pound type beyond 1 000 MW class, the number of thecasings of the low pressure turbine becomes only two byuse of the 3 000 rpm 48 inch blade. Moreover, althoughthe 3 600 rpm 40 inch titanium blade has been used in the60 Hz-700 MW class steam turbine, the 3 600 rpm 40 inchsteel blade, which can be scaled down from the 3 000 rpm48 inch blade, can be applied. The dimension of the cas-ing for the 3 000 rpm 48 inch blade is nearly same as thatfor the conventional 3 000 rpm 40 inch steel blade becausethe outer diameter of the casing can be reduced due to thesmall boss ratio design. And therefore, the total cost ofthe power station including the turbine building can be re-duced remarkably. Table 1 shows the specification of the3 000 rpm 48 inch blade and Fig. 1 shows the overview ofthe 3 000 rpm 48 inch blade.

2. 1 Aerodynamic designIn the aerodynamic design, the flow pattern and the

blade profile are designed. The flow field of the low pres-sure end blades has a large expanding wall inclination dueto the high-expansion ratio. Especially, it becomes verycomplicated as the flow field becomes three-dimensionaland the interaction between the shock wave and the bound-ary layer occurs because of the high Mach number flow.In the aerodynamic design of the 3 000 rpm 48 inch blade,the fully three-dimensional design is employed to reducethe blade loss and to optimize the flow pattern using thethree-dimensional multi-stage viscous flow analysis codesuch as DENTON’s code based on the latest numerical ap-proach.

Considering the radial body forces generated by theblade profile, the stacking, and endwall configuration, de-sign parameters such as degree of reaction, flow angle,and mass flow rate distribution along the blade height are

Table 1 Specification of 3 000 rpm 48 inch blade

Fig. 1 Whole view of 3 000 rpm 48 inch blade

controlled to optimize the flow field. The curved stackingstationary blade (BOW stacking stationary blade) and theshaped end wall (End Wall Contouring) are employed tocontrol the reaction degree of the span wise direction ofthe stage. They reduce the loading of the stationary bladebase and keep the turning angle of the rotating blade assmall as possible. These effects result in the optimal load-ing distribution of the blade and the reduction of the sec-ondary flow losses that are the large parts of total losses inlow pressure turbine. Furthermore, for the high Mach flowregime, the new designed Converge-Diverge passage pro-file with a negative contoured suction surface is employedto control the shock wave intensity in order to reduce theamount of the shock losses. At the same time, it helps toachieve the desirable vibratory characteristics of the blade,

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(a) Longitudinal section

(b) Blade to blade section

Fig. 2 Computational mesh

Fig. 3 Results of the 3-D flow field analysis

of which frequencies tend to descend when you try to keepthe tip passage in a conventional design of a long blade.

The CFD code and the mesh arrangement are welltuned using the results of the cascade tests and the modelturbine tests so as to get enough accuracy of the flow char-acteristic analysis. Figure 2 (a) and (b) show the examplesof the computational mesh arrangement of the longitudinalsection and the blade to blade section for the analysis ofthe last three stage of the low-pressure turbine, and Fig. 3shows the typical result of the analysis. It can be seen thatthere is no deceleration region at blade base and the steamflow is accelerated smoothly and the secondary flow nearthe blade base is reduced enough.

2. 2 Mechanical designIn order to improve the vibratory strength, the

3 000 rpm 48 inch blade adopts the ISB structure that has

(a) Profile (b) Root/Steeple

Fig. 4 Finite element model

the continuous coupled vibratory characteristic at the ratedspeed, getting in contact with the adjacent shroud/stub dueto the untwist deformation. In the mechanical design, it isrequired that the deformation of the blade is predicted ac-curately as well as the static and vibratory stresses. There-fore, three-dimensional finite element analysis technique,which can consider the non-linear effects due to the fric-tion of the shroud/stub and the large deformation, is em-ployed for the analysis of the deformation, the static stressand the vibratory stress. Especially, the large-sized rootand steeple, which makes it possible to reduce a staticlocal stress by about 60%, can be employed for the ISBstructure to improve the strength against SCC (Stress Cor-rosion Cracking), CF (Corrosion Fatigue) and DSS. Thestatic stress of the large-sized root and steeple is analyzedaccurately, taking the non-linear effects into consideration.This code is also calibrated by the measured result of therotating vibration test and the model turbine test. A typicalthree-dimensional finite element model is shown in Fig. 4and the typical result of the static stress analysis is shownin Fig. 5.

To get the maximum structural dumping, numericalsimulation is carried out, changing the geometry and theclearance of the shroud and stub. In addition, the detailedanalysis is performed to check the vibratory strength forhigher modes, using the estimated excitation force whichis obtained from the past model turbine tests, and to tunethe natural frequencies of the lower modes. An exampleof the calculated mode shapes is shown in Fig. 6.

Another resonant stress to be checked in the mechan-ical design is the resonant stress caused by the interactionforce between the blade and vane, because the axial spanbetween the blade and vane is shortened in order to re-duce the dimension of the casing. Therefore, the resonantstress due to the interaction force is evaluated using 3DCFD and 3D FEM. Figure 7 shows 3D CFD mesh of the


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Fig. 5 Result of the static stress analysis

Fig. 6 Result of the blade vibratory analysis

Fig. 7 CFD mesh of 3 000 rpm 48 inch blade and vane

Fig. 8 Examples of panel modes of 3 000 rpm 48 inch blade

last blade (L-0R) and vane (L-0C) of 3 000 rpm 48 inchblade used in the analysis of the nozzle excitation force ofL-0C on L-0R. The ratio of the vane count to the bladecount is 3 to 4, and this ratio is the same as that of theactual machine. The Unstrest ver.14(5) is used as the un-steady CFD code. Figure 8 shows a example of the highermodes calculated by 3D FEM, which is so-called a panelmode. Using these calculated results, the resonant stressof the last blade caused by the interaction force betweenthe blade and vane is predicted, and compared with thatmeasured in the actual loading test described in the latterchapter(6).

3. Verification Test

After the design of the 3 000 rpm 48 inch blade, theefficiency and the reliability of the developed blades areverified by the rotating vibratory test, the model turbinetest and the telemetry test in the field.

3. 1 Rotating vibratory testIn order to confirm the vibratory characteristic at run-

ning speed, the rotating vibratory test is carried out, usingthe high-speed balance facilities in Takasago MachineryWorks of Mitsubishi Heavy Industries, Ltd. After actualblades are assembled into an actual disc, the test rotor isset up in the vacuum chamber of the high-speed balancefacilities. The test rotor is rotated up to 110% over-speed,and the blades are excited by the air-jet to measure the nat-ural frequencies. The natural frequencies of the blades aremeasured by the conventional telemetry system with thestrain gauges and by the non-contact measurement systemwith the optical sensors. The centrifugal stress and thedeformation of the blade are also measured. The blade vi-bratory characteristic is confirmed at not only rated speedbut also the operating speed including low and high cy-cle operation range, as shown in the Campbell diagram.Figures 9 – 11 show the high-speed balance test facilities,


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Fig. 9 High-speed balance test facility

Fig. 10 Rotating vibratory test rotor 48-inch blades

Fig. 11 Results of rotational vibration test by non-contactmeasurement method

the test rotor of 3 000 rpm 48 inch blades, and the test re-sults of vibratory characteristics measured by non-contactmeasurement method, respectively. Figure 12 shows thecomparison of the measured and calculated shroud defor-mation. The deformation of the shroud in rotation is mea-

Fig. 12 Measured and calculated shroud deformation

sured by CCD camera and compared with that calculatedby nonlinear analysis. As shown in Fig. 12, both resultsshow a good agreement.

3. 2 Model turbine testAt the next stage of the verification, both of the effi-

ciency and the reliability of the developed blades are con-firmed using the model turbine test facility. The modelturbine test facility, which can test the steam turbine bladeup to 52-inch low pressure end blade in full scale, hasbeen used for the verification of the developed blade since1986. This facility has contributed to the improvementof the performance and the reliability of a low pressureturbine, carrying out the verification test of newly devel-oped blades under actual and severer operating conditions.Figure 13 shows the model turbine test facility and themain system arrangement, respectively. The maximumsteam flow of this facility is 400 t/h, which is one of thelargest steam test facilities in the world. The conditionof the supplied steam is controlled by a pressure reducingvalve and desuperheater, while the steam flow is controlledby a governing valve and the steam flow rate is measuredby two weight tanks. In order to test the steam turbinein the wide operating range, for the high loading test thatneeds the larger steam flow and the high condenser vac-uum, the turbine output is absorbed by the hydraulic dy-namometer. For the low loading test that needs the smallersteam flow and the low condenser vacuum, the model tur-bine is driven by the driving turbine. A typical operat-ing range is from 30 t/h to 400 t/h of the steam flow rateand from 722 mmHg to 500 mmHg of the condenser vac-uum. In order to verify the vibration characteristics of the3 000 rpm 48 inch blade under actual loading condition,1/1.2 scaled 3 600 rpm 40 inch blade is used.

In the measurement of the vibratory stress, resonantstress was measured in the speed change tests under var-


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(a) Whole view of test facility

(b) Main system diagram

Fig. 13 Actual loading test facility for steam turbine

Fig. 14 Campbell diagram measured in model turbine test

ious loading conditions. As illustrated in the Campbelldiagram in Fig. 14, measured resonant stress of the lowervibration mode was very low because of the large logarith-mic decrement of the ISB structure. Figure 15 shows therandom vibratory stress of the last blade, which was mea-sured in the backpressure change tests under various load-ing conditions. As shown in Fig. 15, the random vibratorystress was also very low and a self-excited vibration suchas the flutter has not been observed in the all tested range.

Fig. 15 Random vibratory stress of the last blades

Fig. 16 Comparison of calculated and measured resonant stressof the panel mode

Figure 16 shows the comparison of the calculated andmeasured resonant stress caused by the interaction forcebetween the blade and vane. It can be said that the pre-dicted resonant stress shows a good agreement with themeasured one and the resonant stress due to the interac-tion force is also small.

From these data, the high efficiency and the high reli-ability of the 3 000 rpm 48 inch have been verified.

3. 3 Field telemetry testAs the final verification test of the 3 000 rpm 48 inch

blade, the field telemetry test is carried out. In the fieldtelemetry test, 6 strain gauges are attached on the bladeand the signal of the vibratory stress measured with thestrain gauge is transmitted to the stationary side by useof the transmitter installed in the balance holes. The vi-bratory stress of the blade is measured under all oper-ating conditions including start-up/shut-down operation,over speed operation, and the various loading conditionfrom the no load to the full load. And also the backpres-sure change tests are carried out under various loadingconditions, in order to verify the random vibration. Al-though it took more than one month to finish all tests dur-ing a trial operation of the plant, no strain gauge is dam-aged and the valuable data can be obtained. Figure 17


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Fig. 17 Blade with strain gauge (After field test)

Fig. 18 Campbell diagram measured during start-up in fieldtelemetry test

Fig. 19 Vibratory stress measured in field telemetry test

shows the 3 000 rpm 48 inch blade with strain gauges andFig. 18 shows the Campbell diagram during start-up. Fig-ure 19 shows the vibratory stress of the blade measuredunder various conditions. Using the gauge factor, which isthe ratio of the vibratory stress at the critical location and

at the strain gauge position and is obtained from three-dimensional finite element analysis, the vibratory stress atthe critical location is evaluated and the minimum safetyfactor of the blade for high cycle fatigue is evaluated.From these results, it can be said that the reliability of the3 000 rpm 48 inch blade is very high, because the vibratorystress measured in the field telemetry test of the actual tur-bine is very small due to large blade damping. And also itcan be confirmed that the vibratory stress measured in thefield telemetry test of the actual turbine is nearly same asthat measured in the model turbine test.

4. Conclusion

The 3 000 rpm 48 inch blade has been developed uti-lizing the latest design technologies to meet the require-ment of the improvement of the efficiency and the reliabil-ity of the power plant. Thermal performance and mechan-ical strength of the developed blade has been confirmed bythe rotational vibration test, the model turbine test and thefield telemetry test. These blades are designed based onmany experiences to develop a longer blade of the steamturbine and the latest technology proven by many verifi-cation tests. New advanced low pressure end blades areapplicable not only for new turbines but also for the exist-ing turbine refurbishment and the application of this newblading is expected to contribute to the progress of powergeneration industries.

References

( 1 ) Watanabe, E., Ohyama, H., Tashiro, H., Kaneko, Y.and Kadoya, Y., High Efficiency and Reliable NewLow Pressure End Integral Shroud Blades, Proceedingsof International Conference on Power Engineering-93(ICOPE-93), Tokyo, (1993), pp.393–399.

( 2 ) Watanabe, E., Ohyama, H., Kaneko, Y. and Miyawaki,T., Development of New Advanced Low Pressure EndBlades for High Efficiency Steam Turbine, JSME Int.J., Ser. B, Vol.45, No.3 (2002), pp.552–557.

( 3 ) Kaneko, Y. and Umemura, S., Vibrational Characteris-tics of Rotating Blade with Mechanical Damper, Pro-ceeding of the 1995 Yokohama International Gas Tur-bine Congress, Yokohama, (1995), pp.191–196.

( 4 ) Kaneko, Y., Mase, M. and Watanabe, E., VibrationAnalysis of Integral Shroud Blade for Steam Turbine,Proceedings of International Conference on PowerEngineering-97 (ICOPE-97), Tokyo, (1997), pp.455–460.

( 5 ) Pullan, G., Denton, J. and Dunkley, M., An Experi-mental and Computational Study of the Formation ofa Streamwise Shed Vortex in a Turbine Stage, ASMEGT-2002-30331, (2002), pp.1–9.

( 6 ) Kaneko, Y., Mori, K. and Tochitani, N., Analysis andMeasurement of Vibratory Stress of Integral ShroudBlade for Steam Turbine, Proceedings of InternationalConference on Power Engineering-03 (ICOPE-03),Kobe, (2003), pp.189–194.


development and veriﬁcation of 3000rpm 48inch integral

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