CA1170842A - Steam cooled turbines - Google Patents

Abstract

ABSTRACT

Steam-cooling with a steam thermal barrier for vanes and blades in a gas generator or power turbine is disclosed. This has particular application to apparatus for generating useful power by reheating gas generator turbine exhaust gases to directly generate power, or alternatively, by reheating this exhaust gas to directly generate power and further reheating superheated steam by heat exchange for combined gas reheat and steam superheating with optional steam reheating as well.
Steam-cooling offers advantages over cooling with air or water at the higher temperatures characteristic of the high temperature turbine technology gas turbines coupled with the reheat steam cycle. Consequently, higher cycle efficiency is obtainable for turbine initial firing temperatures of about 2500°F (1371°C) and 2050°F
(1121°C) reheat temperatures where over-all combined cycle degradation resulting from steam cooling is below about 1%. A turbine having steam-cooled components can operate at a cycle pressure ratio of about 38 without intercooling and can include steam superheating in the reheat combustor. To accomplish steam cooling, steam is introduced into an internal steam distribution plenum in the turbine blading and is then directed, through nozzles, to the external surface of the blading.

Description

~ ~ 7084~
Steam Cooled Turbines Technical Field . . _ This invention relates to steam barrier shielding and cooling of gas turbine and power turbine blading and has particular application to steam cooling of turbine blading in a combined gas turbine and steam turbine cycle including reheating of exhaust gas formed in the first stage of the gas turbine. Steam barrier shielding and cooling of high temperature components permits high temperature operation and provides addition'al work to optimize the reheat combustor location and to produce more power at improved efficiency.
Background of the Invention Research and development is currently being ~irected toward many configurations of power systems involving gas turbines because of the growing awareness of impending world energy shortages. My co-pending Canadian application 337,963 filed October 18, 1979 relates to the need to focus technical attention on the reheat cycle, the reheat gas turbine cycle ancl combined gas reheat and steam reheat cycle disclosed in that application can appreciably increa,se power plant thermal efficiency to appro~imately an over-all 50% efficiency level. The introduction of steam cooling can further increase this efficiency level to 55% and higher.
It has been thought by those knowledgeable in the art that steam injection in a simple-cycle gas turbine operating in a combined-cycle mode would degrade combined-cycle efficiency. Therefore, steam injection for such application has been limited to NOX control. Ilowever, as will be shown when steam is used first as a coolant and then as an injec-tion fluid in an efficient and innovative dm~

~ ~ 7~8~2 ` manner applying an integrated gas/steam nozzle for the nozzle vanes and rotating blades in a reheat gas turbine, heretofore unexpected combined cycle efficiency improvements can be realized through higher firing temperatures.
In apparatus for generating useful power by utilizing a reheat gas turbine and steam turbine cycle, with or without further rehea-ting of super-heated steam, conditions of temperature and pressure producing the relatively high cycle efEiciencies typically involve sufficiently high temperatures and high pressures to necessitate special cooling of gas turbine components.
Conventionally, blade cooling is accomplished with air to extract heat from the blade surface.
The physical properties of air limit its effectiveness as a coolant medium, and impose an upper - limit to the temperatures attainable in a gas turbine.
Moreover, air must be sparingly and very carefully used due to its dearness and its effect on cycle degradation ~0 both with respect to power and efficiency. At all cycle pressure ratios, and particularly the higher cycle pressure ratios encountered in apparatus involving re-heating gas turbine exhaust gases, very little pressure is available to be utilized in the region between the compressor discharge and the nozzle and blade pressures encountered.
The use of steam cooling appears to be an advantageous alternative to other approaches to gas turbine component cooling being studied, researched and developed such as transpiration air cooling, water cooling and alternative materials of construction dm~ \`\ - 2 -I ~ 7~84~
capab]e of .~ithstanding the conditions encountered.
Many additional problems incident to prior approaches to component cooling i.nclude corrosion of the metal surfaces at high temperatures, low-cycle metal fatigue, radiant heat absorption, and first-stage nozzle and blade surface fouling during start-up and operation.
Summar of the Invention Y
. In order to permit increasing of the operating temperatures and pressures at which a gas turbine is to be operated for maximum efficiency the present invention uses steam as a coolant and as a thermal barrier. If steam is so used in a combined reheat gas turbine and steam turbine cycle advantage is taken of the cold reheat steam formed from the steam turbine in order to provide vane and blade cooling for the gas generator first stage, and extracted steam is utilized to provide cooling for second-stage gas generator blading and the power turbine blading.
Furthermore, the invention contemplates use of film cooling employing steam as the coolant, whereby steam can be used as a coolant to advantage without sacri:Eicing blade surface roughness, which can be critical where high Reynolds numbers are encountered.
Accordi.ngly, the present invention is concerned with turbine blading for placement in a flow of gas and for rotating a turbine shaft and constitutes the improvement comprising a first directing means for introducing steam into the interior of the blading, a second directing means capable of directing steam from the i.nterior of the blading to the external surfaces of the blading so that the external surface of the blading is dm~ 3 -~ ~ 7~84~
covered with a thermal barrier of steam.
The present invention is àlso a process concerned with the production of useful power wherein compressed heated gas is produced for contacting turbine blading for rotating a turbine shaft, and the turbine blading is cooled, the process comprising: introducing steam into the interior of the blading, directing steam from the interior of the blading onto the exterior surfaces of the blading, whereby the external surface of the bl~ding is covered with a thermal barrier of steam which insulates the blading against radiant and convection heating.
Brief_Description of the Drawings In the drawings which illustrate embodiments of the invention:
Figure 1 is a schematic view of a gas generator with a reheat gas power turbine driving an electric generator, a]ong with a steam turbine driving a second electric generator in a combined reheat gas power turbine and steam turbine cycle.
Figure 2 is a side elevational view, partly in diagrammatic section, of a gas generator with an associated second reheat combustor.
Figure 3 is a schematic view of a gas generator with a combustor cavity associated with a reheat gas power turbine and provided for steam superheating and steam reheating.
Figure 4 is a longitudinal sectional view of an axial flow reheat combustor cavity for effecting gas turbine reheating, steam superheating and steam reheating.
~igure 4A is a transverse sect~onal view of the co~bustor of Fig. 4~
Fig~re 5 is an enlarged transverse sectional view of a tube for superheating steam and/or reheating in dm:~\\, ~ 4 1 3 7 ~ 2 the combllc;tor cavity of Fiq. 4.
Ili.c3ure 6 is a side el.evational view, partly in diagrammatic sec-tion, of a gas genera-tor and a relleat power turbine, such as can be used in a combined reheat gas power turbine and steam turbine cycle, the components of which can be cooled by steam to permit high temperature and high pressure operation according to the present invention.
Figure 7 is a schematicized sectional view of a first-stage nozzle vane showing regions of typical radiation and convection heat flux around the vane, bounded by combined heat transfer coefficients for 2000F (1093C) turbine inlet temperature.
Figure 8 is a graph of emissivity and absorptivity as a function of the natural logarithm of pressure in PSIA. for steam and for combustion products.
Figure 9 is a transverse sectional view of a first-stage nozzle vane adapted for use of steam as a coolant.
Figure 10 is a sectional view of the trailing end of the nozzl.e vane of E'igure 9, takcn substantially upon a plane passing along section line 10--10 of Fig. 9.
Figure 11 is an enlarged view of the leading edge of the head of the nozzle vane of Fig. 9, showing flow of steam therethrough.
Figure 12 is a lorlgitudinal sectional view of the leading edge of the nozzle of Fig. 11, taken substantially upon a plane passing along section line 12--12 on Fig. 11.
Fig~lre 13 is an enlarged sectional view of the nozzle portions of the vane of Fig. 9, showing velocity profiles for the gas generator first stage dm~ 5 _ I ~ 7~84~
nozzle vane steam nozzle.
Figure 14 is a perspective view of a steam-cooled gas generator first-stage rotating blade.
Figure 15 is a transverse sectional view of the rotating blade of Fig. 14, taken substantially upon a plane passing along section line 15--15 on Fig. 14, and showing circulation patterns of steam therethrough.
Figure 16 is an enlarged view of the front portion of an alternative embodiment of rotating blade or vane incorporating leading edge transpiration steam cooling and thermal barrier steam flow.
Figure 17 is a schematicized sectional view of a gas aenerator incorporating steam cooling.
Figure 18 is a graph showing the relationship of adiabatic expansion efficiency as a function of firing temperature for various methods of cooling.
Figure 19 is a graph showing the specific heat of steam and air as a function of temperature.
Figure 20 is a graph describing the ab.solute viscosity of steam and air as a function of temperature.
Figure 21 is a graph showir.g the conductivity of steam and air as a function of temperature.
Figure 22 is a temperature-entropy diagram for the reheat steam turbine cycle with cooling steam applied to a reheat gas turbine.
Figure 23 is a cooling steam schematic flow diagram for a reheat steam turbine and reheat gas turbine.
Detailed Description of a Preferred Embodiment of the Invention .. _ _ .. .... _ ... _ _ The possibility of use of a combined reheat gas dm~ - 6 -I ~ 7û8~1 ?

turb,i.ne and reheat steam turbine cycle leads to the prospect of using extracted steam from a conventlonal steam cycle for cooling purposes. As aforementioned, such a combined cycle is disclosed in my Canadian application 337,963 and the present invention, th~ugh .laving application to the cooling of turbine components in other than combined cycles, is described in conjunction wi.th this cycl.e. Figures 1 through 5 illustrate this cycle which is fully described in Canadian application 337,963; The cycle utilizes second generation gas turbines now being developed and made available on the market.
, In Figure 1, gas generator 20 receives ambient air through inlet line 22, producing compressed air by compressor 24, which is driven through shaft 26 by gas generator turbine 28, which is powered by gas produced in first combustor 30 to which compressed air is fed through line 32. Fuel enters combustor 30 through fuel line 34. Reheat or second combustor 36 receives high velocity exhaust from gas generator turbine 28 through reheat axial flow annular diffuser 38 and discharges reheated gas through axial flow annular acceleration duct 40 to power turbine 42, which drives first electric generator 44 directly by shaft 46. Reheat gas output leaves power turbine 42 through exit line 48 and passes into heat exchanger 50 prior to discharge through stack line 52. Exit gases through exit line 48 pass through three stages of heat exchanger 50, the first being superheater 54, where superheated steam is produced through superheat line 56: the second being evaporator 58, where water from storage drum 60 and combined recycled dm:`\```" ~ 7 1 3 7~842 water is evaporated in line 62; and the third being economizer 64, where water from storage drum 60 and recycled water from line 66 is heated before being evaporated in line 62 of evaporator 58. In economizer 64, recycled water entering from line 66 is warmed for evaporation in line 59 before entering evaporator 58.
Superheated steam leaving superheater 54 through line 56 enters and drives steam turbine 68, which directly powers second electric generator 70 through shaft 72.
Condensate from steam turbine 68 is formed and collected in condenser 76 and pumped by condenser pump 78 through line 80, along with steam formed directly in turbine 68, through line 82, to heater 86, the output of which is fed by boiler feed pump 88 to recycle line 66. It is to be particularly noted that a tandem arrangement of ga.s generator and turbine combination can be utilized, with first electric generator 44 being powered by two such arrangements, the reheat exhaust gas output of which feeds into exit line 48.
A modified cycle is illustrated in Figure 3 wherein gas generator 20 Eunctions in the same dm:\\`\ - 8 -1 ~7~8~2 manner as the c3as generator described for Fiy. 1, and steam turbine 68, colldenser 76, heater 86, and pumps 78 and 88 function in the same manner as in Fig. 1. However, the output of gas generator 20 enters cavity 90 through axial flow annular di.ffuser 38, cavit~ 90 being, a combined reheat combustor and superheater, reheatiny exhaust gas from diffuser 38 of gas generatox 20 and discharging re-heated gas through a~xial flow annular acceleration duct 40 to drive power turbine 42, which drives first electric generator` 44 by means of shaf.t 46. Cavity 90 also super heats the output from evaporator 58, entering cavity gb through line 92 and leaving through line 56, for driving steam turbi.ne 68. Reheating of steam also cccurs in cavit~
90, the steam entering cavity 90 through line 94 and leavinc3 cavity 90 through line 96 to enter steam turbine 68 at a stage downstream of the inlet from superheated steam line 56. Steam turbine 68 drives second electric generator 70 through shaft 72, and recycling of condensed output through line 66 occurs th;rouqh economizer 64 and evaporator 58.
Fig. 2 is a representa-kion of gas c3enerator 20 in association with reheat combustor 36, although gas generator 20 could also be associated with cavity 90 for performing both a reheat ancd steam superheating function. In either instance, a substantially linear axial flow of gas from gas generator 20 to po~er turbine 42 is maintained. Gas gener-ator 20 is made up of combustor 30 and air compressor 24, which has stages 100, 102 and 104, 106 and 108 of a five-stage low pressure section, as well as s-tages ].10 and 112, wllich are representative stages of a 14-stage high pressur~
section. Comb~lstor 30 discharges compressed hcated air to stages 114 and 116 of a two-stage higll pressure sec~ion P'~ _ 9 _ ~ c~ a~3e 1~.1 7~ c~ t-a~Jc ~
pl^eSSU.re SeC`tiC)II thereof. Shaft 26~ COI-~ICCts corn~re--;sor

2~ with gas ge3lerator turbine 2~. ~ nurnhel- of h:igh temperatllre ancl higll pressure ratio gas tu~bines are now on the market, c3as generator 20 in Fig. 2 illus-tratiny the second generation L~ 5000~del, o-ther model designations curxently available co-.L~ercially including the LM 2500, JT 9. RB-211, Spey and the Mars.
Gas generator~20 in Fig. 2 is coupled wi-th reheat combustor 3~, showing fuel line 35, fuel nozzle 120, annular combustion and acceleration region 122, diffusex 124, and the power turbine 42. It is to be understood that a plurali-ty of fuel nozzles 120, arranged concentri-cally procluce the annula.r flow of reheat gas which ~rives power turbine 42.
Alternatively, the output of gas generator 20.can feed cavity 90 shown in Fig, 4~'4A,' ~or pe~foxming botk a reheat and a superhea-ting func-tion. Air from diffuser 130 passes around struts 132 and is heated in combustion reg;.on 134 as fuel nozzle 136 discharges fuel for combustion therei.n. E'uel ent:ers cavity 90 through fuel line 138 and is ignited by spark plug 140, Superheat and reheat helical tubing coils 142 are shown some~hat 5chematically in ~ig. 4, -having the configuratioll of FicJ. 5 in enlarged detail.
Boiler steam header 1~4 furnishes stearn through line 92 to cavity 90, and after boilér steam has tr.aversed its helical path through ca~ity 90 r it leaves cavi-ty 90 through line 56 for collection in boiler steam header :L~8. Reheat steam leaves cavi.ty 90 thrc)llc3h reheat stea:ll ou-tle-t line 96, enter-iny rehca-t stcalll header 146. ~oiler stcam enters cavity 90 thLough line 92 and headcr 1~ hi].e hoi.:ler steam exits cavity 90 throug]l line 56 and heaclcr 1~8. ~eheat pc~/ - 10 -8 ~ ~
steam enters cavity 90 through reheat steam inlet line 94 from header 150. Insulation 152 surrounds cavity 90, diffuser 130, headers 146, 148, 144 and 150, and also surrounds inlet power turbine acceleration duct 154.
Exit gases pass through inlet nozzle 156 to drive turbine 42, passing next to power turbine first stage nozzle 156 as shown in Fig. 4 O The inside surface 157 of inlet duct 154 is in the shape of a nose of a bulletl while outside bell mouth 196 has the shape shown. Fig. 5 shows stream-lined fabricatea superheat tubing 142 containing perforatedairfoil shaped extended surface sheath 158 surrounding three tube sections 160, 162 and 164 for enclosing superheated and/or reheated steam, the respective diameters of the tube sections varying according to the flow rates and throughput demanded of each component.
A portion of the hot gas flows in and out of the perforations to heat the steam. The sheathing provides streamlined sections for gas directional control. The spacing between coils provides pressure drop and mass flow control.
If steam is extracted from a conventional steam cycle for cooling purposes~ the above-described combined cycle can optimize the gas turbine-cycle pressure ratio, while at the same time taking advantage of the exhaust steam from the high pressure steam turbine section generated b~ the dm~

8 4 ~
steam t~bine for first-stage vane and blade cooling of thé gas generator. Extracted steam a-t a lower pressure can also be used for cooling the second-stage gas generator blading and the power turbine blading. Under typical reheat steam turbine steam conditions of 2400 psig-1000/1000F(16.56 MPa and 538/538C), such as the conditions disclosed in U.S. Patent 4,314,442 to Rice issued February 9, 1982, U.S. Patent 4,272~953 to Rice issued June 16, 1981, and U.S. Patent 4,384,452 to Rice issued May 24, 1983, steam posæ sses superior physical characteristics as a coolant when substituted for conventionally used air. With use of conventional methods of nozzle and blade casting, together with metallurgical techniques within the present state-of-the-art, base-load firing temperatures of the gas turbine can be increased according to the methods and apparatus of the present invention frem the present level of about 2100F for`baseload operation (1149C) to about 2400F (1315C) or higher. These physical characteristics, including specific heat, viscosity, conductivity, radiation absorptivity, and Prandtl nu~ber, lead to substantial advantages of steam for cooling, even taking into account the specific volume and sonic velocity. The conventional coolant, aix, must be used sparingly and very carefully, primaxily due to its effect on cycle degradation with respect to power and efficiency. Nevertheless, despite the fact that a very small pressure drop exists between the compressor discharge and the nozzle and blade, and despite the fact that at higher cycle pressure ratios, the coolant temperature increases to its detriment; highly developed convection and film-cooling techniques have been developed using air as a o~olant, initially for aixcraft use, and later for use with industrial gas turbines.
Turbine nozzle vane and rotating blade design components can be steam cooled and thermal barrier protected to achieve dm:i\\ - 12 -i~7~
about ~5~ ],O~`Jer 11eating Value (LI~V) comb.i.l1ed c~cle eficiencies according to the teachings ~1hich will be hereafter outlined.
Fig. 6 shows a gas genexator 4l~, gas generator blading 43~, a reheat combustor 30A, lncluding fuel line 32A, fuel nozzle 34~, annular co~bustion region 36A, diffuser 38~, and power turbine 40A and power turbine blading 42A. Ordinarily, a plurality of fuel nozzles 34~
are arranged concentrically to produce the annular flow of reheat gas to drive the power turbine located riyhtwardly of Fig. 6. .Such gas gener~tor, reheat combustor and power turbine com-prises components, such as blading 42A and 43A, which mus~
withstand the temperatures generated by heated yases in annular gas generator combustion region 35A and powex turbine combustion region 36A. Accordingly, the temper-ature of combustion gases obtainable is limited by the ability of internal components to main.tain structural inte-.
grity and be free from corrosion for prolonged periods of operation. With the present invention, cooling of these components by use of steam perm.its h:iaher operatin~ lemper--atures to be attained with minimum cycle degradatlon and with greater efficiency than if such components were air cooled. An over-all increase in efficiency can result over and ~x~.e the power plant ef~icienc~ approachmg or exceeding 50% L~
(Lower ~.eatin~ V.alue) attainable with the ~o~er plant fully disclosed in the aforementioned Canadian application 337,963. In fact, an efficiency ap roachina or exceeding 55% LHV is possible with use of steam cooling a~d thermal barrier protection as taugnt by the present in~ntion.

~rom calculations the theoretical efficiency obtain-able can be shown to be 56.7% L11V. This optimum efficiency does not comPensate for blade cooling, bu.t indicates the theoretical maximum obtainable. If steam is used for blade cooling and if the steam can be heated to n'/ - 13 -1 ~ 7~84~
~it:h;.I~ 300i` (167C) of the fi~-in(J teJn~crature i2I t:hc process of blanket cooling or Lhermal barrier cvo].ing, then the combined cycle efficiency ~70u:ld he slightly degraded to a value of about 56o L~IV, and if a safety factor is applied for the.steam (to be heated to a lo-,rer value of approximately 600F (333C), then 55~ LHV can reasonably be selected as a target efficiency value.
Considerable incentive exIsts to maximize blade cooling and thermal barrier protection from a cycle efficieney standpoint by use of steam. Cycle degradatIon is greate;r with liquid water as a coolan-t, and it is possible and likely that higher firiny temperatures offer little or no advantage over someW~at lowex fi.rmg temperat.ur.es usin~ steam as a eoolant.
~odern-day gas turbines which fire at elevated temperatures and at higher cycle pressures encounter greater heat flux to the combustor and blading components, including far more radiant heat transfer as well as greater eonveetive heat transfer. With respeet to the eombustion liner, flame radiation is a dominan-t factor.
A number of variables are involved, including flame temperature, combustion pressure, flcime luminosiL-y, flame dimensions, combustor geometry and the emissivity and absorptivity of both the gases and the liner~ Com-bustion eonditions, such as type of fuel, development of soot or smoke, and the like, effect the radiation produced from the three primary sources, namely, non-luminous CO2, non-luminous water vapor and hot ~lowin~
sub-mi.cron carbon particles, oft:en described in the art as radian-t clouds. Radiant heat transfer for known first-generatioIl 70 psia (~82G50 Pa) gas turbiIle liners P'~3/

I 1 7 ~
have been rep~rted to be up~ar(ls of 175,000 BTI~ ~>er Sq. Ft.-Ilr. (1984500 KJ/M -llr. ). ~ second-generation 25 ra-tio combustor, on the other hand, has a radiant heat transfer rate of 230,000 to 280,000 BTU per Sq. Ft.-Hr.
(2608100 to 3175200 KJ/M 2-~Irr 1). At these rates and a -1500F (833C) temperature differential between the flame and liner, the effective heat transfer coefficient is about 167 BTU per Sq. Ft -Hr, per ~F (3~10 KJ/M -llr. -C 3.
Expressed in terms of average firins temperature, this co-efficient ~ecomes abou-t 400 BTU per Sq. ~t.-llr. per F
(8168 KJ/M -Hr. -C ). Fig. 7 ~illustrates typical heat flux values, where the total transfer coefficients around vane 4~ are resolved into radiation and convection values, which are independent of each other. Leading edge 50A of vane 48A points toward the combustox flame for most gas turbines. Hiyh ratio second generation yas turbine absorb radiant heat at a level of about 400 BTU per Sq.
Ft.-~lr. per F (8168 KJ/M -Hr. -DC ) the heat flux being related to inlet temperat:ure and nc)t to II.lme temperature.
The leading edgc gas film is squee~.ed to a very thin film duè-to the stagnation pressure and velocity profile character-istics, which factors cause a very high convection heat transfer rate. The total transfer coeEficient is the sum Gf the two terms, shown or. Fig. 7 ancl is 740 ~ per Sq. Ft.~ r.
per F (15110 KJ/M 2-Hr l-C~l) Heat transfer by radia-tion varies as the fourth power of the absolute temperature in accordance with Stefan-Boltzmann's Law. Fig. 8 provicles values o emissivity and absorptivity as a function of the natural logarithm of

3~ pressure in PSI~. These values can be applied to the ahove mentioned radiatioI- law to calculate radiant heat flu~;.

15 ~~

1 :~ 7~ 2 T}l~? pressure side of vane 48A has a rather constant heat transfer COeL l-icient of about 160 BTU per Sq. Ft.-Hr. per ~F (3267 KJ/M -Hr. 1 _oC~l) as shown in Fiq. 7. However, the suction side develops a hump two-thirds of the way downstream, at which point the boundary layer is thought to thicken and become more turbulent. At this point, the heat transfer coefficient doubles to about 320 BTU
per Sq. Ft.-Hr. per ZF (6534 KJ/M Z-Hr~-l -oC-l) The radiant heat portion of the total heat transfer coefficient is reduced where the blade surface is exposed only to the radiation of the carbon particle cloud and not the CO2 and H2O. This portion might be about 25% of the total for a gas temperature of around 2100F (1149C).
In conventional air-cooled first stage nozzles for gas turbines constructed according to present technology, many cooling weep holes on the vane surface and leading edge are provided to permit cooling air to exit outwardly at a low velocity to be accelerated by the stream and carried over the vane nozzle surface. The main gas stream must accelerate the coolant and in the process only a partial film is formed due to the mixing process that takes place and gaps between the exit ho]es commonly referred to as weep of gill holes.
If steam is substituted for air, first-stage nozzle design of the gas generator is modified to the form shown in Figs. 9, 10, 11 and 12.
Fig. 9 suggests that leading edge 56A of vane 54A be cons-tructed separately from vane body 58A and allowed to float as leading edge 56A expands radially inwardly from a fixed peripheral anchor point.
Leading edge 56A is aligned to vane body 58A by alignment surfaces 60A.
It should be noted that leading edge 56A receives the greatest heat flux from radiation and convection, and separation of leading edge 56A will keep heat away from body 58A of vane 54A.

dm: ~ ~\' - 16 -1 ~ 7~84~
I,c;l<lillg e(l9e wcep-holes c:an c]clditio~ ]ly be cml,lo~cd on .~ad:ing edge 5~,~ to assist in a~)SOI-~ti-)ll of rad.;.anl heat.
Built-in steanl nozzles il~ the form of full }en~Jth slots are incorporated in leading edge 56~, as is hest seen in the enlarged view of Fig 11- Leading edge 56A
can be made of solid ceramic material or coated with ceramic thermal heat barrier 62A and body portion 58~
has extended surface horizontal fins 64~, hest seen in cross-sectional view Fig. 12, for convective heat re-moval. Fins 64A also provide steam-nozzle guides or ducts. Floating leading edge 56A can be made removable from body portion 58A so that xeplacement or recoating is easy during overhaul of vane 54A. Preferably, steam exits along the entire surfaces of vane 54~ through lead-ing edge nozzles 66A tangentially~ or as closely as possible to tangentially, with a very small reactive velocity.
The steam exit velocity through nozzles 66A should be equal to or slightly higher than the entrance velocity of the gas stream. An average velocity near the leading eage might be about 700 feet per second ~13 m per second). Steam can also exit through trailing edge nozzle 67A Oll trailing edge portion 69A and will preferably emerge with a velocit:y nearly equal to the gas velocity near nozzle 67A which is typicail~ a~out 2000 ft/sec (610 M/sec) for such turbines Velocity profiles of the steam and gas are shown in Fig. 1~ at the nozzle slot exits. The laminar films are shown on the vane surface. The steam nozzle Reynolds number, ReS, relative to diameter is about 2.7 x 10 whereas the main gas stream Reynolds nun~er, Reg, relative to diamet~r is about 2O5 x 10 as shown.
The amalgamation of the nozz~e steam velocity pro~ile, both at the leading edge nozzle and at the trailing edge nozzle is apparent. Where the steam and gas meet laminar flo~

ms/

I ~ 7Q~,42 :`~ i '. ~ ~; r(~- l)o~ CJc~ c~ o ~ l i rlcl c~ f c~ t- o~
l~c-~ tJake all(l a smooth transit:iol~l of velociLy l)eL~Jeen thc two fluids unti~ the stcam b~ul~c3aL-y la~cl~ is es~ablished.
rhe pressllre inside tile nozzle cap~ made possib]e b~7 the selcctioll o the ~as turbine cyclc-pressure ratio of about 38, as well as by using the hiclh pressure steam turbine exit line to -the reheater is about 60 psi. (.414 MPa) greater than the surrounding static gas pressure when allQwance is made for the pressure drop in the steam line to tne internal vane steam forward nozzles. This pressure differential then provides ah exit velocity of about 700 feet per second (Z13 M/sec). The desired result is for the leading edge steam tangential velocity to match t~e gas stream veloc~ty in order to distur~ the flo;7 as little as possible, and in the process,a blanket or boundary layer of steam ~an be established over the surfaces of vane 54A to serve both as a radiant heat abosorbant and a retardant to conduction of heat to the blading.
A thin laminar flow is maintained on the surfaces of vane 54A similar to the oil film wedtJe oL a journal bearing. This boundary layer shoulcl be sta~le on the pres~sure sides of the vane (concave) surface 68A, but on the suction (convex) side 70A, there is a thickenin~
of the boundaryl~yer where film eEfectiveness migh-t be reduced. Therefore, horizonta] extended surface fins 72~ (best seen in cross-sectional view Pi~. 10) form trailinq edae nozzle 67~ for internal convection coolina. Appropriate steam. turbine e~traction pressure is ~Ised to prov:ide the necessary trailin~ ed~e velocity of approximately 2000 ft/sec. (~)10 ~sec). ~mple pressure is available from the steam turbine. Stcam mass flow to the trailina edcle can reasorlably be assumcd to 111 'i /

<~ ^o~ t~ , of t~l~ r~ ; f~ C;t~lln :~;:i.t velocity ~;11 be about e~ ? to tllc no:~zle e~i t veloc;ty to rcduce e~it loss~s.
The s-team nozz]e edge should be clS s~larp as prac-tical to minimize eddy current disturbanccs at the nozzle e~cit point near nozzle 66,-~. The widih o the nozzle should preferably be rather small and is calculated to be about 0.010 inches ~.254 mm) for 1-1/2% steam flow and one inch radial distance between individual nozzle vanes at the leading edge. However, a 60 psi (.414 MPa) pressure drop should ~eep the slots blown clean and prevent plugging by ash deposits.
Nozzle side-wall cooling is also important when hic~h inlet temperatures are involved. Side slots and ~eep-holes on both the outer and inner walls are necessary to film-cool the surfaces with steam.
The general arranc3ement shown in Pigs. 9 to 12 points to other advantages of steam cooling. The steam can be used to start the gas generator as -there is ample flow of steam. Considering a conventional LM
5000 gas generator and 1wl/2~ of t:he gas flow as steam, the rate of steam aeneration required is about 15,000 lbs. per hours (6804 Kg/hr). At a water feed rate of 20 lbs./KWHr (9.07 Kg/KWHr) the power made available would be about 1005 HP (750KW). With blade-cooling steam, the starting horsepower would be about twice this value.
Another advantage of steam cooling is the nozzle and blade-cleaning ability of s-team .hen t}-e gas turbine is shut down. Steam at about 30 pSi2 ( . 207 MPa) can reach sonic velocity out nozzles 66A, ~lly ash deposit can be cleaned off the bladincJ. A built-in form of steam lance is accordincJly pres~nt.

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cl S~.eca~ i.S a~ d to t:})e xc)l:.L:ir~cr ~l-lde:;
~e gelle~a:l collclpL of cool;ng ~; ';U~CIc'; t~ f~r t}-ie no~
va-les can hc u~ed such as ~ladc 80~ i]lustrated in ~ig. 14 in a perspec-tive pictorial vie~l and sho-~r in section in Fig. 15. OverlappillcJ nozzle slots 82~ are substituted for the single loncr nozzle slo-t of the b~ade of Fig. 9 to provide blade 80~ added strength to with-stand centrifugal forces. Nozzles slots 86A are used at the trailin~ edge of blade 80A for the same reason.
It is impor-tant tha-t ~he air foil surface be as clean as possible. Air cooling as presen-tly used produces surEace rou~hness which is detrimental such surface roughening being capable of being reduced considerably particularly with respect to the leading edge 88A if adequate convective cooling can be realized with steam. Weep-holes (not shown) in leading edge 88A are to be avoi.ded if possible for aerodynamic reasons; however such holes could be '! helpful in reducing the radiant and convection heat absorption but in any event should be used sparingl~
if at all.
A second;form of rotating blade or stat~onary vane ada~ted for steam cooling is sllown in Fig. 16 where anotller alt.e`rnative to the high leading edge heat flux involves incorporating transpiration cooling at a sacrifice in surface roughness. Blade 90A shows how such a process can be accomp].ished using transpiration ~ 20 -C~OC~ t ~ rr~ t~l t~ c~c)o l ~l n ~: . ', t ~ i s 1 ~ 7 ~ 8 4 rneteL-~c~ f:rol,l plenurrl c~amber !)21~ to trar~spiration - ch~ambers 9~,?~ through srnall or;~ice~s 96A. The steam under controlled pressure is then allowed to ~low througll porous wire braiding 9~. Transpiration cooling using air as the coolant is presently under active investigation, as i.s well-~nown by those s~ Led in the art, with extensive research and development work presently under way. However, attention has not yet been focused on the advantage of steam for transpiration cooling or to t~e ~uestion of whether extraction steam is purer than extracted air~ When appropriately controlled vaporizati.on processes are used, the steam introduced into plenum chanlber 92A is purer than air presently used, and deposits which would otherwise ~orm or plug the pores in wire braiding 98A are less with the steam than with the conventionai air cooling.
Transpiration cooling can also be applied if re~uired to the rear suction side of blade 90A, here greater turbulence and greater heat :LU~.
exist, such method again beiny at the sacrifice of .
surface roughness anc3 its effect on heat transfer and aerodynam;cs~
The cooling steam is introduced through the t~rbine wheel in a manner similar to that i.n which air is introduced with present conventional air cooling technic~ues. However, in order to obtain a higher pressure inside the blacles, a modified approach is desirable, as shown in Fiy. 17. Stearn enters steam header 102~ of yas generator 100~ Gas feed line lO~A furnishes combustible c~as to fuel nozzle I 1 7~84~
106A, and heatcd gases are formed in combustion region 108A to pass through turbine wheel llOA. Cooling steam is introduced through turbine wheel llOA where steam lines are brought through bearing struts 112A to a labyrinth packing or seal ll~A at the small diarneter of shafting 116A. Steam then flows internally in annular cavity 118A
up to entrance holes 120A of turbine wheel llOA. No internal vanes in the cavity are present in order to minimize pumping losses and steam-head buildup. The blades are preferably sealed off with flexible metallic-type gaskets at the roots of the blades so that steam pressure can be maintained. In addition, power turbine blading also must be cooled in view of the 2000DF (1093DC) inlet temperature of gases contacting the power turbine blading. Although the same approach to cooling can be used as for cooling turbine wheel llOA, the power turbine blades are larger and have a greater surface area, necessitating careful utilization of steam in order to avoid degrading cycle efficiency. Low-pressure extraction steam is used for this purpose, in a quantity oE approximately 1-1/2~ of the main gas stream.
' Small orificies 250 in the inner wall of annular cone ; shaped chamber of 118A near the exit to turbine wheel llOA allow a small portion of steam to impinge on turbine wheel llOA surface to provide disc cooling.
A significant advantage in using steam as a coolant comes in thé area of low-cycle fatigue and thermal shock experienced by metal nozz]e vane and blade components during start-up. More importantly, however, low-cycle fatigue and thermal shock resistance are important considerations on full-load trip-out or load dumping from full load to no load. The effect of low-cycle fatigue is perhaps ten times more deleterious on emergency shut-down than on dm~ - 22 -~ 7~
s~ r~ t~c cc~ r~r~;sc~r clic;ci~.lr~ air ueoches tl)c hot vanerj and blades a~ cJas tcmperatur~
drops ;uddenly. ~ith IS';' of stcam as a coolant wheL-e ~ small start-up boiler and a steam accuJnolator - 5 for shut-down are provided, it is ~ossible to proyram the cooling steam on start-up and shut-do~n.
Proper control valves and process controllers enable use of such a start-up boiler and stearn accumulator.
Ater shut-down, steam is admitted at a programmed rate of decay with built~in time constants to keep the bladeis under a controlled temperature drop.
A further consideration with use of steam concerns hot corrosion (sulfidation), which becomes more critical at the el~vated temperatures encountered in the proposed process. Various coa'~ings have been developed to help protect blade surfaces where corrosive products of combustion come in direct contact with the blades and the protec~ive coatings.
I~ a blanket of steam is rnaintained over the blade surfaces, contact of corrosive elernents ;s .su)~stan--tiall~r if no~ entirely, avoidcd at the blade surfaces.
Only the heavy particles of corrosive cornbustion products can penetrate the steam film to reach the blade surfaces because of their mass and momentuJn.
Leadirly ed~e~ 88~ or 56~, however, are exposed and protective coatings would be required, such as protective coating 62~ sho~;n in Fi~. 11. Thermal--barrier ceramics and/or plasma-sprayed coatings can ~e applied, such as alloys of platinum, chromiurn, and aluminum.
Th~ (>re~c)ing approaches and technic~u~s enable full advantage to be ta~en of the superi()r 23 ~

~ ~ 7 ~
~ c~i ~,r ~.it~ ; c~ (lc~ ~()c)l.lllt clr~(l th(rJ(~
_~a~ ier, alld (~nclb]e intcc$rdti.o~l of t;he coo~.ant w:ith the steam cycle. 'lhere are dic;tinct di.fferellces in the physical propertics of steam and ai.r which affect heat transfer and cooling characteristics of c;as turbine vanes, blades and discs. It is necessary to provide for steam cooling due to the unsuitabi]ity of eonventional air eooling techniques at higher firing temperatures. For example, Fig. 18provides an indi-eation of the problem faced with current teehno]og~
t~lereadiabatic expansion effieieney is plotted as a funetion of firing temperature for various methods of eooling. The eurve presents a broad band of effieieney levels, and shows that there is a point of diminishing return in terms of eyele effieiency and gas turbine output with use of air as a coolant, further suggestinc3 that it is possible that this point has now been reached for industrial gas turbines at about 2100F
(].149~C).
Perhaps the most significant differenee between steam and air as a eoolant eoneerns thei.r relative speeifie heats, or the capaeity to absorb heat for a given temperature rise. Fig. 19 graphieally compares the speeifie heat of steam with air and shows the distinct difference and advantage of steam over air. While the speeifie heat of air remains almost eons-tant over a wide range from 500F
to 2400F (260C to 1316C), the specific heat of steam, besides beinc3 about twiee that of air, a].so fluetuates with both pre~ssure and temperature. The specifie heat of s-team in the re~lion of eoncern for gas turbine blade eoolinc~ varics from about 0.85 at ~ 2a -~,c,/

. . . ` I 1 7~84~
~00 ~->si (~ ) clr~(l 500~ noC) t~ ct V~31U~' 0~
boul: 0.63 at 100 p~i (0.690 ~.~a) and 2000~1 ~lO~C).
~he specific heat rises frorn .50 to . 60 at lesser pressures for the same temperature ran~e, again reflecting over twice the heat absorbing capacit~ of air and ma~ing steam a ~ar better coolant for gas turbines than air.
Further, steam has a lower viscosit~ than air in the range of pressure and temperature encoun-1~ tered in blade cooling. Fig; 20 shows the relation-ship ~etween absolute viscosity and temperature for both steam and air. ~he viscosity of both air and steam rises ~Jith temperature at about the same rate.
Although the viscosity of steam is only moderately lower than air, there is an advantage due to friction factors and Reynolds numbers for heat transfer and flow characteristics in the laminar film region and boundary layer near the contact surface, the Reynold,s number being inversely proportional to the viscosity~
Another characteristic of st~am relevallt to heat transfer is i~s thermal conductivity, na~lely, ~he ability of heat to be transmitted within or through the fluid itself. Fig. 21 sho~s the relation between thermal conductivity and te~perature for both steam and air. The superior characteristic of steam is particularly apparent at the elevated temperatures encountered in blade cooling, the conductivity of steam being about 50C~ higher than air at 1500F (816~C).
3n T~adiation to and from non-lurninous ~ases is another important factor to b~ considered in heatin~

_ 25 -~ 1 7~84~
cl~c~ c~ (J ~)I` b].~lc~ (J, i~clrt:ic~ c~ (vc~t~(l ~m~ u~? :L~ L.~ o~ ~Irc)u~ld ~ '.r~ (1.31~C). ~t such tem~eratures~ 30-~O or more of the hcat ~ransfer is in tlle form of radi.ant enercly. Whe~n a thin layer of relatively cool steam blankets the vane~s and blades, the steam is an excellent raciant hea-t absorher.
At elevated temperatures, with carbon dioxide and including : water of combustion, and water introduced for nitrogen oxides control purposes (both-the carbon dioxide and water bein~ present in low partial pressures), the :Eilm ~f pure steam around the vane and blades will absorb much of the combustion radiant heat, and will re-emit this energy at a much lower energy content and heat rate.
:Much of the radiant heat w.ill be carried away with the continuous flow of steam before it can enter the blades.
Cooling air presently used for cooling does not have :.this characteristic, and almost all of the radiant heat present passes ~irectly into the blades.
A further consideratlon in e~aluating heat transfer is the Prandtl number, a higher va.lue indicatin~ significantly better heat transfer, as is well-known in the ar-t. For steam at low pressures, this number drops from 0.98 at 600F (316C) to 0.88 at 1200F (6fi9C), while for air the corresponding Prandtl number is 0.70 at 600F (316C) and 0.65 at 1400F (760C).
Other comparisons are also significant in evaluating the relative comparison of steam and air involve -the specific volume and sonic velocity relative to pipin~ and passageway sizes and critical area noz~le cons:ideratiol-ls.

- 2~ -r!l.s /

I ) 7~84 ~rl~ o~ c~ Or ~ st~r~ /c~r~ ir fluid of ~he s-,~ea~n turbirle :in tl~e c~mhil~e(l cycle ;.s g;vell in .Fiy. 22 where cycle temperatur(-:s are plotted agclinst cycle entropies. Irracing the steam turbine cycle, water is pumped from poirlt 130A throuqh the economizer where it is heatecl to point 132A the evaporator entrance point. The water is evapora-ted from point 132~ to point 134A in the heat-recovery boiler and then receives its first superheat from point 134 : 10 to point 136A in the second gas turbi.ne comhustor by steam coils 142, ~ig. 4. The steam then expands through the steam turbine to point 138A providing useful work. The steam is again superheated from point 138A
to point 140A in th.e second combustor by coils 142 before expandinc3 again in the steam turbine from point 140A to point 142A, the condensation point.
The extracted steam that is used to cool th~
reheat gas turbine blading is hypothetically separated from the main gas stream for analysis as will now be discussed wi.th reference to Fig. 23.
Steam is extracted at the high pressure steam tur,bine exhaust line to tlle reheater cold reheat polnt 138~ and is heated in the ~as generator first stage nozzles and blading to point 144A some 300 to 600F
(167 to 333C) below the first combustor exit temperature.
The steam at this point is part of the gas turbine flow.
It can be noted that the gas turbine cycle ratio of 38 allows for 60 to 70 psi steam pressure drop. About 10 psi (68.95 KPa) would be need~d to t.ransport the steam to the bladir~J -makinc3 ~bout 50 psi (3~5 KPa) avail-ab:l.e for ~ooling which is much more pressure than no~
realized for air cooling.

~ 27 -Jll'; /

I ~ 7 ~
l,Xt~ l(`t:iC)~ cll~l~ 110~ ';}10~`/11 i~l i'i(J 22 i~;
i~ewiC;e used for coolirlg thc second-st~ c c3as genercllor bladin~3 Steam at point 15~ is extracted for coo]iny - the first-sta(Je }~laclil-g oF the power turbine ~7herc the~
steam is hea-ted from point 158~ to point 152~ and exparlded to point 154~. It should be noted t,hat the steam previously used for Lhe gas genexator is heated in the second combustor all the way to point 148~ the full reheat temperature befo~e expansion as it is a homogeneous mixture. The gas turbine reheat process increases the powex deveioped by the gas generator cooling st'eam.
~eat is then recovered in the heat recovery boiler from points 150~ and 154A to poin-t 156A, the assumed 300 ~ ~149 C) stack temperature. In accordance with the calculated steam cycle heat balance, the heat in the exhaust steam is converted to useful power at an efficiency of 40.43'~ L~V to complete the cycle.
The steam flow can be broken down fox purposes of analysis schematically as 5hown in ~ig. 23, The cooling steam is separa-ted from the main s-treal-n in a hyp,othetical way but no-t in reali-ty a procedure oE-ten done in analyzing single and double automatic extraction/
admission steam turbine cycles.
In the case of the reheat gas turbine, steam in line 160A is extracted from the high pressure steam turbine exhaust line 162A and is heated in the first staye gas aenerator blading - designated as heat exchanger 164~. The steam is now part oE the'~as turbine flow. Incremental fuel is introduced in line 163~ and burned in combustor 170~ to preheat tl-;s steam as it passes through and around the blading as will be shown later. The main fuel for t:he gas turbine ms~

I 1 7~8~2 enters t.h~.c~ 31-1 i n:l.ct ~ i.ne ~ A.
'I`ile coclillc,,l s~tealn, artCl^ e~ nd:ing throuc]h gas c3enercltor turbine 174A, i.s again heated in comblls~or 176A at heat exchange 178A, the heat be;ng suppliecl as fuel in line 180A. Fuel for tlle main steam stream enters from line 182A; -the cooling steam for the power turbine from line 184A and tlle main gas turbine air flow from line 186A.
The heated ga,s ~enerator turbine cooling ' 10 steam then expands through power turbine 188A to produce power after which the steam is ex~austed to heat recovery boiler heat exchanger l90A, where heat is extracted to heat the feed water line 192A. The final stack temperature of 300~ (149C) is assumed to be compatible with the main heat-recovery boiler stack temperature.
Cooling steam for the power turbine in line ! 194A is extracted from the low-pressure steam turbine 196A and is heated by power turbine first-stage blading designated as heat exchanger 19~A. The steam now part of the gas turbine stream then continues to expa,nd to the exhaust where it is also recovered in heat recovery boiler heat exchanger l90A~
Extraction steam that flows in lines 200A~
202A and 204A is used to heat the feed water from 101 F (38.3 C) to 240 F (116 C). Boiler feed pump 206A raises the temperature to 250 ~ (121 C). It is assumed that the make-up water arrives at an en-thalpy (h) of 70 to be compatible wi-th the heat balarice cycle of l~'ig. 23.
The total cooling steam is no-~ hea-ted in pselldo heat exchanger 20~A by ma}ie believe fuel frorn line 210A to brinc3 the steam bac~i to the initial 2~15 - 2~ -m~;/

~ 1 7~84 ~
~ic~ (~h.5S r~ l0~ (53~ ollclit-ic)~ {~
~challcJcr 20~ is uscd to cleteL-mi,lle t:he Luel r~quired to comple-te -the loop so tha~: the coolinc3 s-team-cyc]c eff:i,ciency can he determilled and evalua-ted.
Fig. 23 can also be used for the simple-cycle gas turbine by elimina-t:ing combustor 17G~ and by extracting the steam at the appropriate steam pressure to allow for adequate injection pressure and piping , pressure drop.
Using data developed, Table I summarizes the eEfeet steam substitu-tion has on over-all eombined-eyele effieieney. This table indieates that the over-all eombined eyele will be slightly degraded by about 0.7% '0.4~ Points). It is doubtful, however, if 2400/2000 F (1316/1093C, levels for base-load operation ean be aeeomplished without resorting to steam eooling or some other type of cooling beyond ! air cooling in current use today.
A signifieant inerease in output of 36.2% ean be noted when eompared to the original gas turbine work. This ealeulates to be 6 -o increase ~or each 1%
ste~m. This addecl bonus should improve the installed cost per unit of power output of the eombinea eyele of this invention.
It ean be noted in Table ~ that the eooling steam develops more gas-generatox turbine work than the eooling air it displaees. The cooling steam work is 21.16 first-stage plus 9.66 second-stage equals 30.82 BTU/lb of air for steam cool:ing.' The eooliny air wor~ is 10.59 first-staye plus 4.01 seeond-staye equals 14.G0 BTU/lb of air for eooling air.
The gain by difference is 16.22 B'rU/lb of air. This extra work made available by the steam raises the xeheat r!' _./

I ~ 7~84 2 ~ t; ~ c;~ t~ ~ (>ViCi~ c~ J})~l~ t~ t~tcll -~ las generatc)r tllrbine wor~ deve1c)l}~xl ,mus~ equal t:he wor~ rcquire(l-to compress the air whicl-l xemains constant.
'rhe rehea-t pressure thus will rise accordinc31y. The cooling steam flow can be increase~ beyond the amount needed for blade cooling to optimize the split of total expansion ratio between the gas-generato~ turbine and the power turbine.
It can also be noted tha-t the cyc~e eficiency o the steam is tabulated in Tab:le I- to be 43.21 percent which is appreciably above the 40.43 percent level of the associated condensing steam turbine and accounts for the improvement in combined cycle eiciency over what could be expected using air as a coolant if it were possible to do so at a 2400~ turbine inlet temperature.

~lS/

Ct~ t ~itllnut (`oo] in~ _ St~ rn Cooli ----~--~
Air SLeain Worlc Fuel ~k)rk F~lel z %~TU/I,b Air ~ Ai r~ U/I.~ t ir_ B~UlLb ~ir F irst-St;lge GG 8 3 GT 10.5~ 10.81 21.16 . 60.~U
ST , ~ 17 _ -~;ecolld-Stage GG ~ 1. 5 GT . I~.01 4.29 9.66 28.85 . . . _ 3 . 0 9 GT .8G . 4.50 27~34 ST 4. S2* - 5. 89 Cycle Eff % 19.92 15.1 ';0~47 116.79 .. _ . . . _ . . . .. _ .
Change in PT Exit Temp ~F l~2 Original GT Work ~TU¦Lb ~ir 206~8S
Credit for Exit Temp Inc in GT Work % ; 24.62 @ 40~ uork BTU¦Lb ~.52* Inc in Total ~ork~Vs GT ~1Ork)% 33.64 ~ncrease in GT WQrk BTU¦Lb Air 50.72 Original Comb Cycle Eff ~D_' . 56.00 Increase in ST Work BTU/Lh Air lg.67 0ver-all Cycle Eff X LHV
To~al Xncrease in Work BTUILB 7D.39 50~ GT & 50% ST 55.62 Total Fuel BTU/Lb Air 131.89 0verall Cycle ~ff % L~
Over-all Ef of Tnc ~ork ~ L}IV 53.37 60% GT & ~0% ST 55 *-~.ssuming stea~ ~eing heated to 300 F (1670 C~ of ffle ls~ ana 2nd cc~b~stor ~e~peratures. This approach is not very c-{itical and does not change the over-all cycle efficiency very greatly.
oeations: GG ~ Gas Generaeor GT = Gas Turbine PT = Power Turbine ST = Steam Turbine Gas Turbine Cycle Pressure Ratio = 38 First Combustor Firing,Temp ~F - 2500 (1371 C) (Xnc1udes fuel added for stealll he~tin~
Second Combustor Firing Temp F - 2050 (ll21 C) ~Xncludes fuel added for steam he?.ting).

,.

-- 32. --i ~ 70842 ~rllel-~ c:c~ Ld bc~ o~e pe~ ancl conc~ rltt:ion ~n usi.rlg steam as a blade cool.allt in thclt abo~t 20'~
(o the stcam flow) make--up coul.d l~e required to replace the loss in s-team out the stacl.
Abbrevi.ations used throughout the application are as follows:
.j HP ~ Horsepower MPa - Megapascals E~ - Expansion Ratio LHV - Lower Heating Value h - Enthalpy SIU conversion factors are as follows:
BTU = 1.055 K~
Bl'U/lb - 2.326 KJ/Kg PSI = 6~895 KPa ~ = 9/5C + 32.
; Throughout the specification and claims, unless otherwise specified, cycle efficiencies.are expressed in terms of fuel lower heating value abbrevi.ated(LHV), temperatures ~re given in de~re~s Fahrenhcit, pressures in pounds per square inch absolute, costs in U.S. dollars, power in kilowatts (KW), energy in British Thermal Vnits (BTU), and parts and proportions in percent by weight.
The foregoing is considered as illustrative only o the principles of the invention. Further, since numerous modifications and changes will readily occur to those skilled in the art, it is not desired to limit the invention to the exact construction and operation shown and described, and accordingly, all suitable modifications and equivalents may be resorted to, falling within the scope of the inventi.on. Similarly, 3~
Jns/

I17~
as the state of the art of gas turbi~es advances ~: through improved higher temperature metallur~y, higher temperature construction materials, hi~her firing temperatures, higher compressor efficiency, higher turbine efficiency, and higher compressor ratios, the general temperature and pressure relation-ships between the first combustor, second combustor, gas generator and power turbine are considered to move upward accordingly from the levels presented here-tofore and fall within the scope of the invention.
~ his application is a division of Canadian application 337,963 filed ~ctober 18, 1979.

dm~ 34 -

Claims (47)

THE EMBODIMENTS OF THE INVENTION IN WHICH AN EXCLUSIVE
PROPERTY OR PRIVILEGE IS CLAIMED ARE DEFINED AS FOLLOWS:

1. In turbine blading for placement in a flow of gas and for rotating a turbine shaft, the improvement comprising: a first directing means for introducing steam into the interior of said blading, a second directing means capable of directing steam from the interior of said blading to the external surfaces of said blading so that the external surface of said blading is covered with a thermal barrier of steam.

2. A gas turbine comprising turbine blading for rotating a turbine shaft during the expansion of gas through said turbine, a supply of steam, a first directing means for introducing steam from said supply into the interior of said blading, a second directing means for directing said steam from the interior of said blading to the external surfaces of said blading whereby the external surface of said blading is covered with a thermal barrier of steam which insulates the blading against radiant and convection heating.

3. The gas turbine of claim 2, wherein said blading comprises stationary vanes and rotating blades, said first directing means introducing steam from said supply into the interior of said vanes or said rotating blades or both.

4. The gas turbine of claim 3, wherein said blading includes an internal steam distribution plenum for receiving said steam introduced from said supply and said second steam directing means comprises at least one steam nozzle communicating with said plenum and open to the external surface of said blading.

5. The gas turbine of claim 4, wherein said blading includes a leading edge portion and a trailing edge portion, said steam nozzle being placed in said leading edge portion and directing steam across said external surface toward said trailing edge portion, said trailing edge portion of said blading including therein an internal coolant steam distribution plenum, a third steam directing means introducing steam from said steam supply to said coolant steam distribution plenum, a second steam nozzle placed at the trailing end of said blading, said second steam nozzle communicating with said coolant steam distribution plenum in said trailing edge portion.

6. The gas turbine of claim 5, wherein said steam nozzle placed in the leading edge portion of said blading includes a plurality of heat exchange fins extending from the interior surface thereof to provide convection cooling of said leading edge portion, said heat exchange fins being parallel to the flow of said steam from said internal steam distribution plenum to said steam nozzle opening for guiding said steam through said nozzle.

7. The gas turbine of claim 5, wherein said coolant steam nozzle on said trailing end of said blading includes a plurality of heat exchange fins and ducts therebetween placed on the internal surface thereof wherein coolant steam flowing therethrough serves to absorb heat from the suction side of said trailing edge portion of said blading.

8. The gas turbine of claim 5, wherein said blading includes a stationary vane, said vane comprising said leading edge portion and a body portion which includes said trailing edge portion, said leading edge portion being attached at a fixed periphery anchor point to said body portion, said leading edge portion being separated from said body portion to form a cavity, said leading edge portion being allowed to float and expand radially inwardly from said fixed peripheral anchor point, whereby heat is kept away from said body portion.

9. The gas turbine of claim 8, wherein said steam nozzle placed in the leading edge portion of said vane includes a plurality of heat exchange fins extending from the interior surface thereof to provide convection cooling of said leading edge portion, said heat exchange fins being parallel to the flow of said steam from said internal steam distribution plenum to the nozzle opening for guiding said steam through said nozzle, said coolant steam nozzle on said trailing end of said vane including a plurality of heat exchange fins and ducts therebetween placed on the internal surface thereof, wherein coolant steam flowing therethrough serves to absorb heat from the suction side of said trailing edge portion of said vane.

10. The gas turbine of claim 8, wherein said leading edge portion on said vane comprises an external protective plasma-sprayed metal alloy coating selected from the group consisting of platinum chromium and aluminum.

11. The gas turbine of claim 8, wherein said leading edge portion on said nozzle is provided with weep holes for allowing steam to issue therefrom from said internal steam distribution plenum, whereby the influx of radiant and convection heat on the leading edge portion can be substantially reduced.

12. The gas turbine of claim 11, wherein said leading edge portion on said vane comprises an external coating of a thermal barrier ceramic.

13. The gas turbine of claim 8, wherein said leading edge portion of said vane is removable and replaceable.

14. The gas turbine of claim 8, wherein said leading edge portion and said body portion of said vane are provided with a smooth surface, whereby aerodynamic efficiency is improved.

15. The gas turbine of claim 8, wherein said leading edge nozzle in said stationary vane directs steam from the interior of said vane tangentially onto the external surface of said body portion, and wherein said trailing end nozzle of said vane directs coolant steam from said coolant steam distribution plenum parallel to the gas flow along said trailing edge portion.

16. The gas turbine of claim 5, wherein said at least one steam nozzle placed on said leading edge portion of said blading comprises a plurality of steam nozzles which form two rows placed on opposite sides of said leading edge portion.

17. The gas turbine of claim 5, wherein said leading edge portion on said blading is provided with weep holes for providing steam to issue therefrom said internal steam distribution plenum, said leading edge portion being provided with a woven wire covering placed over said weep holes whereby transpiration cooling of the leading edge portion takes place and the influx of radiant and con-vection heat on the leading edge portion is substantially reduced.

18. The gas turbine of claim 16, wherein said blading comprises station-ary vanes and wherein each said row consists of one or more steam nozzles.

19. The gas turbine of claim 16, wherein said blading comprises rotating blades and said steam nozzles placed on the leading edge portion of said rotating blades comprise a pair of rows of parallel staggered slots open to the exterior surface of said blades, the rows of slots being placed on opposite sides of said leading edge portion and extending substantially across the full width of said rotating blades, whereby steam exiting from said nozzles provides a blanket of steam over the external surfaces of said trailing edge portion.

20. The gas turbine of claim 17, wherein said blading comprises rotating blades.

21. The gas turbine of claim 3, wherein said blading comprises rotating blades, said supply of steam comprises a steam header and said first steam directing means comprises a first steam passage from said steam header to a labyrinth packing at the small diameter of said turbine shaft, a second steam passage from said packing to entrance holes positioned within a turbine wheel which supports said rotating blades and a third steam passage from said turbine wheel for introducing steam from said turbine wheel into the interior of said rotating blades.

22. The gas turbine of claim 21, wherein said rotating blades are sealed with flexible metallic-type gaskets between the blade roots and the turbine wheel rim whereby steam pressure can be maintained.

23. The gas turbine of claim 5, wherein said leading edge steam nozzle is positioned so as to direct steam from the interior of said blading tangentially onto the exterior surface of said body portion.

24. The gas turbine of claim 21, wherein said second steam passage comprises an annular cavity connecting said labyrinth packing and the entrance hole positioned within said turbine wheel, said annular cavity being provided with orifices for cooling the internal stationary parts and first-stage discs and shafting of said turbine.

25. A power production system comprising a first gas turbine for driving a compressor, means to reheat gas expanded in said first gas turbine, a power turbine receiving said reheated gas, said first gas turbine and said power turbine including blading to rotate respective turbine shafts, heat exchange means for producing steam from the gases exiting said power turbine, a steam turbine, means to direct said produced steam through said steam turbine, means to direct steam exiting said steam turbine into the interior of said blading of said first gas turbine or said power turbine or both, means to direct steam from the interior of said blading onto the external surfaces of said blading whereby the external surface of said blading is covered with a thermal barrier of steam which insulates the blading against radiant and convection heating.

26. Turbine blading for placement in a flow of gas and for rotating a turbine shaft comprising: a leading edge portion and a trailing edge portion, a first steam distribution plenum in the interior of said leading edge portion, a second steam distribution plenum in the trailing edge portion, means to introduce steam into the interior of said blading, at least one steam nozzle placed in said leading edge portion and communicating with said first steam distribution plenum, said steam nozzle capable of directing steam across the external surfaces of said trailing edge portion and a steam nozzle placed at the trailing end of said blading.

27. The turbine blading of claim 26, wherein said blading comprising a stationary vane or a rotating blade.

28. In the production of useful power wherein compressed heated gas is produced for contacting turbine blading for rotating a turbine shaft, and said turbine blading is cooled, the improvement comprising: introducing steam into the interior of said blading, directing steam from the interior of said blading onto the exterior surfaces of said blading, whereby the external surface of said blading is covered with a thermal barrier of steam which insulates the blading against radiant and convection heating.

29. In the production of useful power wherein compressed heated gas is produced for contacting turbine blading for rotating a turbine shaft, and said turbine blading is cooled, the improvement comprising: introducing steam into the interior of said blading, directing steam from the interior of said blading substantially tangentially onto the exterior surfaces of said blading, the relative velocity of said steam directed onto the surface of said blading being substantially zero with respect to the velocity of said compressed heated gas contacting said blading, whereby the external surface of said blading is covered with a thermal barrier of steam which insulates the blading against radiant and convection heating.

30. The process of claim 29 wherein said blading comprises stationary vanes and rotating blades, said steam being introduced into the interior of said vanes or said rotating blades or both.

31. The process of claim 30 wherein the production of useful power comprises a combined reheat gas turbine and steam turbine cycle wherein said compressed heated gas passes through a gas turbine and is reheated in a combustion cavity for driving a power turbine, said gas after leaving said power turbine exchanging heat to form steam, said formed steam being introduced into the interior of the blading of either or both of said gas turbine and power turbine to provide said thermal blanket of steam.

32. The process of claim 29 wherein said blading includes an internal steam distribution plenum and at least one steam nozzle communicating with said plenum and open to the external surface of said blading wherein steam is introduced into said plenum and directed out through said steam nozzle to blanket the external surface of said blading with said thermal barrier of steam.

33. The process of claim 32 wherein said blading includes a leading edge portion and a trailing edge portion, said steam nozzle being placed in said leading edge portion and directing steam across said external surfaces toward said trailing edge portion, said trailing edge portion of said blading including therein an internal coolant steam distribution plenum wherein coolant steam introduced therein is also distributed through said coolant steam plenum to a nozzle placed at the trailing end of said blading whereby said trailing edge steam velocity is substantially equal to the main steam velocity.

34. The process of claim 33 wherein said steam nozzle placed in the leading edge portion of said blading includes a plurality of heat exchange fins extending from the interior surface thereof to provide convection cooling of said leading edge portion, said heat exchange fins being parallel to the flow of said steam from said distribution plenum to the nozzle opening for guiding said steam through said nozzle.

35. The process of claim 34 wherein said leading edge portion of said stationary vane is attached to a body portion which includes said trailing edge portion, said leading edge portion being separated from the body portion to form a cavity, whereby heat is kept away from said body portion and said leading edge portion is allowed to float, said leading edge portion containing said steam distribution plenum.

36. The process of claim 33 wherein said steam nozzle on said trailing edge portion of said blading includes a plurality of heat exchange fins and ducts therebetween placed on the internal surface thereof wherein coolant steam flowing therethrough serves to absorb heat from the suction side of said trailing edge portion of said blading.

37. The process of claim 34 wherein said leading edge portion on said blading comprises an external protective plasma-sprayed metal alloy coating selected from the group consisting of platinum, chromium and aluminum.

38. The process of claim 37 wherein said leading edge portion on said blading is provided with weep holes for allowing steam to issue therefrom from said distribution plenum, whereby the influx of radiant and convection heat on the leading edge portion is substantially reduced.

39. The process of claim 38 wherein said leading edge portion on said stationary vane comprises an external coating of a thermal barrier ceramic.

40. The process of claim 33 wherein steam nozzles placed on the leading edge portions of said blading are in the form of two rows placed on opposite side of said leading edge portion, said steam issuing from said nozzles providing a thermal blanket of steam on the concave and convex surfaces of said blading.

41. The process of claim 40 wherein the steam nozzles placed on the leading edge of said rotating blades comprise a pair of rows of parallel staggered slots open to the exterior surface of said blades, the rows of slots placed on opposite sides of said leading edge portion and extending substantially across the full width of said blade, whereby steam exiting from said nozzles provide a blanket of steam over the external surfaces of said body portion.

42. The process of claim 38 wherein the leading edge portion of said blading is provided with a woven wire covering placed over said weep holes whereby transpiration cooling of the leading edge portion takes place.

43. The process of claim 32 wherein said thermal blanket of steam is provided by exhaust steam from said steam turbine.

44. The process of claim 40 wherein the amount of steam used to blanket said stationary vanes and rotating blades is about 6% by weight of said compressed heated gas passing through said combustion cavity and said power turbine.

45. The process of claim 33 wherein said turbine is started by introducing steam into the leading edge portion of said blading and introducing steam into the trailing edge portion of said blading, allowing steam from said leading edge portion and said trailing edge portion of said blading to strike said rotating blades to heat said blades and start rotation of said blades.

46. The process of claim 29 wherein said steam is applied before start-up to sonic clean the blading whereby soot, ash and other deposits are removed.

47. The process of claim 43 whereby cleaning is performed while the turbine is running.