US3475108A

US3475108A - Blade structure for turbines

Info

Publication number: US3475108A
Application number: US705465A
Authority: US
Inventors: Walemar Zickuhr
Original assignee: Siemens AG
Current assignee: Siemens AG
Priority date: 1968-02-14
Filing date: 1968-02-14
Publication date: 1969-10-28
Anticipated expiration: 1986-10-28

Description

Oct; 28, 1969 WJ ICKUHR 3,475,108

BLADE STRUCTURE FOR TURBINES v I Filed Feb. 14, 1968 (5 Sheets-Sheet 1 I W. 'ZICKUHR BLADE STRUCTURE FOR TURBINES Oct. 28, 1969 Filed Feb. 14, 1968 I l l 3 Sheets-Sheet 2 Fig.5a

Fig.5b

| l i v I 1 1 l 6 I i Fig.5c

l I l I 1 a E Fig.5d

Oct. 28, 1969 w. ZICKUHR 3,475,108

BLADE STRUCTURE FOR TURBINES Filed Feb. 14, 1968 3 SheetsSheet :5

[ 1 kWh v (25% m v ado AD [mm] 0 160 zbn Fig.7 "sm .u 11 Q 11 11 1 Pr3 Pr Fig.8

United States Patent US. Cl. 415194 14 Claims ABSTRACT OF THE DISCLOSURE A blade structure for turbines such as high-pressure axial turbines. The turbine has at least One ring of stationary blades and a rotor which carries a coac'ting ring of rotary blades, these blades forming at least one of a series of turbine stages connected one next to the other in the direction in which the pressure of the turbine-driving pressure fluid drops, while this fluid expands during driving of the rotor. The blades have at a mean diameter at a turbine region where a given operating factor prevails and at the rated capacity of the turbine stage, with a Reynolds number of at least 2x10 an opening value e=e/ t which is at least equal to 0.5 as an average value for the blades where e is the smallest distance between adjoining blades and t is the distance between corresponding points of a pair of adjoining blades. This operating factor at the latter turbine region is either a pressure which is equal to or greater than 2 atmospheres absolute or a region of the turbine where there is an increase of not more than in the specific volume of each stage.

My invention relates to turbines and in particular to blade structure for turbines.

Turbines such as steam turbines are known in the most widely different forms. In their basic construction they include at least a stationary ring of blades and a rotor which carries a ring of blades for rotation with respect to the stationary ring of blades. The adjoining stationary and rotary rings of blades form generally at least one of a series of turbine stages located one after the other in the direction in which the pressure of the steam or other pressure medium drops, while this pressure fluid expands to drive the rotor. 1n the design and dimensioning of such turbines it is always attempted to make use of the strength of the materials in conjunction with the greatest possible operating safety while maintaining construction costs as low as possible and achieving at the same time a good efficiency.

It is an object of my invention also to achieve these objectives, which as to say to provide a turbine, particularly a steam turbine, which is contrasted with conventional steam turbines will have an improved efficiency of operation and at the same time a reduction in the construction costs.

My invention is based upon recognition of the fact that these objectives can be achieved by a proper selection of the primary turbine dimensions, such as, for example, the shaft diameter, the critical rotary speed at a given operating speed as well as by the proper selection of the opening value e=e/ t for the blades. In this latter equation t represents the circumferential distance between corresponding points of a pair of adjoining blades, which is to say the blade distribution of a given ring of blades, while e represents the smallest distance between a pair of adjoining blades at the mean entrance diameter of the operating fluid. The term mean entrance diameter is explained further below.

In particular my invention is based upon recognition of the fact that the above-mentioned opening value of the blades, is of particular significance with respect to the best possible operating efficiency and construction cost of the turbine. Thus, it is a primary object of my invention to ice construct the blades in such a way and to provide them with such an opening value that the result will be an improvement in the operating efficiency together with a reduction in the construction costs.

Up to the present time it has been customary to provide at the mean blade diameter an opening value 6 which at a maximum is 0.4 when a particular turbine stage is designed at its rated capacity to operate at pressures equal to or greater than 2 atmospheres absolute. It is to be understood that the mean blade diameter, or admission diameter, signifies the mean diameter of that turbine stage where the pressure drop is equal to the mean value of the pressure drops of the several stages of a given turbine casing. The mean admission diameter is approximately equal to the inner stream dimension plus the blade height. Thus, Traupel states in his book Thermische Turbomachinen (1st volume, Springer Publishing Co., 1958) at page 174, that for a high pressure turbine where the degree of reaction r=0.5, the best value for the angle a (entrance angle of the absolute entrance speed 0 of the steam, which at a degree of reaction of 0.5 equals the exit angle 5 of the relative steam exit speed W2) should be 17". On page 330 of the same book, Traupel points out that at a high pressure region of a steam turbine small volumes are accompanied by small angles. Appropriate design data for the optimum points of blades follow for a conventional constant pressure stage for an entrance angle a =14 and for an exit angle 5 :20"; for a conventional high pressure stage a =fi =17 30', which corresponds to an opening value e=e/t of approximately 0.3, since e=e/t is approximately equal to the sine of the angle 8 These recommendations for an extremely small opening value have clearly their basis in the fact that it is assumed that an increase in the opening value must result in an unacceptable decrease of the circumferential operating efficiency and because it is assumed that with large opening values the blades are very short and as a result the necessarily accompanying increase in the blade tip loss 1' (flow loss at the blade tips) introduces a further efficiency reducing factor.

In another known text on turbine construction (Fliigel, Die Dampfturbinen 1931, page 190) it is recommended that normally with drum stages fgzx =fgfl between 0.25 and 0.40 should be selected. Furthermore, Fliigel writes that with sharp volume increases during progressive eX- pansion, particularly in low pressure regions it is essential to resort to steeper exit inclinations so that there will be no excessively rapid increase in blade length which would be undesirable for the flow, and thus at high pressure stages tga =O.60 to 0.70 is used.

In another book on turbine construction (Bauer: Der Schiffsmaschinebau, 1927, pages 65-67 as well as and 151) there is given only for the last stage of condensing turbines tgfl-values between 0.45 and 1.5, where opening values of approximately 0.40 to 0.72 are present.

According to my invention it has been recognized that an increase of the opening value e/t above 0.4 is desirable not only in the last stages of low pressure regions with large volume increases of approximately 200% and more at each stage, but also an improvement in the operating efliciency and a reduction in the construction costs can be achieved in a most unexpected manner with an increased opening value for high and intermediate pressure regions of turbines where there is only a volume increase of approximately 3% from stage to stage at the high pressure region of the turbine up to a volume increase of 15% from stage to stage at the intermediate pressure or mean pressure region of the turbine.

Thus, it is an object of my invention to provide blading for steam turbines and the'like where coacting rings of stationary and rotary blades form at least one of a series of successive turbine stages in the direction of pressure drop of the rotor driving the steam. My invention resides in providing, at pressures at least equal to 2 atmospheres absolute or at an increase in the specific volume from stage to stage of less than 15%, at the rated capacity of the turbine stage, and with a Reynolds number an opening value e e/l at the means blade diameter which is equal to or greater than 0.45 as a mean value for the stationary and rotary blades, where e=the smallest distance between adjoining blades,

t=the distance between corresponding points of a pair of blades,

w =the relative exit speed,

b=the blade width, and

1/*=the kinematic viscosity.

The requirement that the Reynolds number be at least equal to or greater than 2 10 is based upon the fact that the increase in the deflection impulse of a grid shovel, which is to say with a given deflection the increase in the amount of flow throughput and thus the increase in the opening value, is comparable with the increase in the lift coefiicient of an aeroplane wing. Such an increase in the lift coefiicient is only possible if the drag of the primary flow stream at the boundary layer, which is to say the Reynolds number, is sufiiciently great. In this event there will be the required impulse exchange between the primary flow and the boundary layer, and separation or breaking down of the boundary layer cannot take place. Developments carried out in connection with my invention have shown, in harmony with these considerations, that a blade profile having a large relative resistance moment, which is large enough for the stage load, and a large opening value can only be achieved above a predetermined minimum Reynolds number, nomely 2 10 at the rated capacity.

My invention has recognized the fact that as a result of the features of my invention the height of the blades, both for the stationary and rotary blades, can be reduced. Inasmuch as the centrifugal force stresses at the root of the blades for the rotary blades is reduced in this way, the possible bending stresses for a given rotary blade increases, which is to say, however, that the blade width as seen in the axial direction and thus the axial dimensions of the entire blading can be reduced. From this factor it follows that inspite of the required increase in the number of stages and cost of a unit of axial length for the purpose of maintaining constant the critical speed, there is nevertheless a decrease in the construction cost on the order of 520% as compared to conventional turbines and blading. Together with this decrease in the construction costs there is an appreciable increase in efiiciency which is on the order of l*5% of the total operating efficiency as compared to conventional turbines, which is to say at least of the theoretically possible increase in efiiciency.

Thus, it becomes possible with the features of my invention to provide smaller blades, to reduce the construction costs of components such as shafts, casings, blades, and at the same time to achieve an increase in efficiency.

On the basis of flow tests it has been determined that the blade tip loss in a first approximation is proportional to the ratio profile width blade length My invention has, therefore, recognized that with full use of the maximum possible axial blade length with a given shaft diameter, taking into consideration the critical speed, the blade Widthwith geometrically similar profiles-can be reduced with an increase in the opening value 2/! in such a way that the ratio b/l=profile width to blade length can be maintained smaller and thus the blade tip loss can be reduced.

More accurate tests, referred to in greater detail below and taking into consideration the gap, piston, and blade tip loss as well as the influence of Parsons number and the opening value on the efficiency, show that depending upon the extent of gap loss and the proportion of centrifugal force stressing to the entire blade stressing, the turbine efiiciency is at a maximum when the opening value, or the opening ratio as a mean value for the stationary and rotary blades, is between 0.50 and 0.75. In the region of greater opening values, which are greater than the preerred region between 0.65 and 0.7, there is a limit where the increase in efficiency can no longer be achieved, but even at these greater opening values there is also an increase in the savings of construction costs, so that the designer can decide in this area on a compromise between reducing construction costs and increasing efiiciency in accordance with which of these advantages he considers to be 'of greater importance. Therefore it is possible that under certain conditions it will be desirable to exceed an opening value of 0.7.

My invention can be used advantageously for turbines having one or more casings, whether the turbines are condensation turbines or the reaction type of back pressure turbines. My invention makes it highly desirable to use high pressure stages as well as to be used in the high pressure regions of steam turbines where a mean pressure on the order of atmospheres absolute prevails. My invention furthermore can be favorably used with steam turbines where there is a mean pressure at intermediate pressure stages on the order of 8 atmospheres absolute. With multi-casing turbines it is possible to use the features of my invention only for the high pressure part, only for the intermediate pressure part, or for both of these parts of the turbine.

A preferred construction of the blading of my invention can be used with advantage with steam turbines where there is an extent of reaction of between 25 and 65% (reaction turbines or turbines with mixed features where there is a compromise between reaction and impulse turbines), without however being limited to these particular types of turbines. A reaction or high pressure turbine is for a given construction cost, especially with large outputs, superior to a constant pressure turbine by approximately 1-1.5 in efliciency.

According to a further feature of my invention the blading is used in steam turbine where the extent of reaction is approximately 50%, which is to say between 0.45 and 0.55 (reaction or high pressure turbine in the narrower sense). Such blading is of advantage particularly because, as shown by speed diagrams, the entrance and exit flow angles of the blading and thus the profiles for the stationary rotary blades are approximately the same. This simplifies the manufacture and provides a uniform load on the ring of blades as seen in the direction in which the pressure within the turbine drops. With steam turbines where there is a reaction extent different from 0.5, it is preferred to define the mean opening value sine 01 sine a; e z

It is also possible to use the features of my invention with advantage for this type of turbine, as mentioned above. Steam turbines suitable for using the blading of my invention are preferably axial turbines, which is to say all turbines where stationary and rotary blade rings follow alternatively one after the other along the axis of the turbine rotor.

In accordance with a further feature of my invention, the blades are constructed in such a way that the relative maximum profile thickness of the blades d /L is between 0.3 and 0.4, where d is the diameter of the blade profile at the inner circuit at its thickest region and L is the chord length of the profile (the longest dimension or the cross-sectional blade profile). This feature provides a large strength for the blades because of the high resistance moments, and in addition within this region there is the greatest efiiciency.

A further increase in efliciency can in particular be achieved with blading having appreciable centrifugal force stresses for the rotary blades, especially for intermediate pressure regions of turbines, by providing the rotary blades with a tapered construction. Such a blade construction differs from a cylindrical-radial blade where the cross section of the blade is constant over the entire blade height and the center of gravity is situated at a radius which extends perpendicularly from the rotary axis, in that the blades taper and become of gradually smaller cross section toward their outer tips. Such an expedient results in a smaller width for the ring of rotary blades and thus a decrease in the gap, piston, peripheral, and discharge losses. An increase in the flow losses does not take place. With constant relative profile thicknesses at the blade root the tapered blades will have a reduction in flow loss at the intermediate regions of the blades, since the relative profile thickness of the blade cross section becomes gradually smaller as the cross section becomes more distant from the blade root. With tapered blade configurations it is possible to combine within certain limits the advantages of relatively thick and relatively thin profiles, which is to say small blades and small flow losses.

It is even better according to a further embodiment of my invention to provide rotary blades which are twisted as well as tapered, so that it becomes possible in this way to better adapt the profile at the individual blade sections to the particular flow angle, and thus select an even smaller profile thickness. Also, the extent of taper, which is to say the reduction in centrifugal force stresses as compared to cylindrical-radial blade forms, is in this case greater than the warped blade forms of taperedtwisted construction. The strength and flow advantages of the tapered-non-twisted and the tapered-twisted blade forms can also be incorporated to a large extent in the rotary blades.

Further features and advantages of my invention as well as the manner of operation thereof are described below in connection with several embodiments and diagrams illustrated in the drawings which accompany and form part of this application and in which:

FIG. 1 is schematic simplified illustration of a developed portion of blade rings with stationary or rotary blades;

FIG. 2 is a schematic illustration of the section of the blades of high pressure turbine of the axial type where the blades of one rotary ring fixed to the rotor are shown in conjunction with axially adjacent blades of a pair of adjacent stationary rings fixedly carried by the turbine casing;

FIG. 3, which corresponds to the illustration of FIG. 1, serves to explain the speed diagram of FIG. 4 and to derive the opening value e=e/ t, where a ring of rotary blades and a ring of stationary blades are shown adjacent each other in the axial direction;

FIG. 4 is a speed diagram in which the entrance and exit angles are also enlarged in accordance with the enlarged opening angle of my invention;

FIGS. 5a-5d are a series of diagrams for explaining the technical advances achieved with my invention, and in all the diagrams 5a5d the pertinent magnitudes have been normalized into a dimensionless form depending upon the opening value e shown at the abscissae, this normalizing being achieved by referring all magnitudes to a corresponding value where e=O.3, which for magnitudes having dimensions is indicated by a dash, and in particular FIG. 5a shows a curve of the total efiiciency of the turbine A17=Zv;

FIG. 5b shows the curve of individual losses, and thus the piston loss 1 the gap loss 11 the blade tip loss r and the e-dependant loss;

FIG. 50 shows the increase in the permissible bending stress fl'Bi and the corresponding or normalized stage number Z;

FIG. 5d shows the drop of the corresponding blade height h, the corresponding active length L which is to say the part of the turbine shaft where the blading is situated, the drop of the mean blade diameter 5, which is analogous to the'curve h, the drop of the corresponding blade width b as well as the ratio o'F/zr which is to say the ratio determined by the centrifugal force at a given stress in tension of the blade root with respect to the permissible tension stress at the blade root;

FIG. 6 shows, on the basis of the operating efficiency improvement according to FIG. 5a, the attainable improvement in heat consumption in a diagram depending upon the increase in the weight diameter in the second casing of the turbine, which is to say at its intermediate or mean pressure casing. In this case the extent of improvement of heat consumption indicated at the ordinate in kcal./kwh. signifies the reduction of the required generating heat for the steam per kwh. output of electrical power, and the weight diameter, indicated on the abscissa and shown with D MD, signifies the diameter of a shaft section of the axial length of the first turbine stage whose weight is equal to the weight of the shaft and rotary blade ring present in the first stage. This imaginary magnitude of weight diameter has proved to be of advantage for turbine calculations, particularly for large turbines, where beside shaft bending stresses the tangential stresses and the influence of the temperature field in the shaft to determine the best possible flow dimensions are also of significance. The use of the diameter as a magnitude influencing the shaft stressing is also proper in connection with the significance of the radial dimensions of the casing for evaluation of operating safety (temperature stressing at casing walls and partial gap flanges). FIG. 6 shows a pair of curved groups, in the one case with the critical speed n =0.8 A=const. and in the other case with n =A=const. In both groups of curves there is with increasing rotary speeds I' which is explained further below, and for increasing mean opening values according to sin ot =0.3, 0.4, 0.5 and 0.6 an indication of the improvement in heat consumption. In this case the curves are shown for a high pressure turbine of the axial type having an extent of reaction of 50% and another with the same mean entrance angle a and means exit angle FIG. 7 is a diagram of the improvement in heat consumption with two other variations of blade construction depending upon the increase in weight diameter and thus the lower curve of FIG. 7 shows the improvement in heat consumption for a blade of cylindrical-radial profile and the upper curve the improvement in the consumption for a blade having a tapered-twisted profile; and

FIG. 8 schematically illustrate-s the corresponding profile configurations.

Referring now to the drawings, FIG. 1 shows the section of blades for an axial turbine of the high pressure type provided with stationary or rotary blades L, L, where the stationary blades are to be considered as uniformly distributed about the inner peripheral surface of an unillustrated turbine easing or the rotary blades are to be considered as uniformly distributed about the exterior periphery of a turbine rotor. The distance e indi cates, as indicated, the smallest distance between a pair of adjoining blades and the distance 13 indicates the dis tance between corresponding points of a pair of adjacent blades. The

radial boundary planes

1 and 2 show where the axially adjoining next ring of rotary or stationary blades is situated, while d is the diameter of the circle which is characteristic of the maximum relative profile thickness and L is the profile chord length (the longest dimension of the blade profile).

The simplified mean or meridian sectional illustration of FIG. 2 shows the rotary blades L of a rotarv rin of blades and the adjoining axially displaced stationary blade rings L in both directions for a pair of stationary blade rings. The stationary rings of blades L are fixed to a casing which is not further illustrated or to a blade carrier 3 at the blade roots 4 of the stationary blades, while the rotary blades L are fixed at their roots or feet 5 to a rotor 6 which is not further illustrated. Thus, as is apparent from FIG. 2, the structure shown is a high pressure turbine of the axial drum type. The height or longitudinal length l of the blades is indicated in FIG. 2 as well as the profile width 11 thereof in the axial direction.

FIG. 3 shows, in correspondence with FIG. 1, a stationary ring of blades L and an axially adjoining rotary ring of blades L, both rings being shown in section only with a pair of blades. The blades have entrance edges 7 and exit edges 8. The steam enters, as is known, from the ring of stationary blades L with an absolute entrance speed c and at a corresponding entrance angle as well as with a relative entrance speed W and the corresponding entrance angle ,6 into the rotary ring of blades L to give up at the latter a part of its kinetic energy so as to drive the rotary ring of blades and the rotor, to achieve in this Way the rotary speed u. As is known the steam expands in the stationary ring of blades L from the initial pressure 2 up to the final pressure 11 which in accordance with i, the s-diagram of enthelpy difierence i -i =H corresponds to and provides an isentropic or theoretical entrance speed c /2gH where g indicates the acceleration of gravity. In accordance with whether the structure forms part of a high pressure or reaction turbine or a constant pressure or impulse turbine, the steam will expand more or less also in the rotary blades L, which is to say it expands from the pressure p to the pressure p corresponding to an enthalpy differential of i -i =H Whether the structure forms part of a high pressure or a constant pressure turbine, it has a characteristic for the latter factors the extent of reaction.

, =i=ra L'+ L H s where H is the drop from stage to stage (the drop from a stationary ring of blades to the adjoining rotary ring of blades). With high pressure turbines there is conventionally a degree of reaction on the order of r=50%, for example, while with action turbines there is, for example, a reaction extent of r=%. With action turbines there is thus a reversal of pressure at steam speeds encountered primarily in the nozzle rings. Thus, the invention is not limited to the magnitude of the degree of reaction, although it is of particular advantage when used with high pressure turbines having an extent of reaction on the order of 50%.

Returning now to FIG. 3 after the steam has given up part of its kinetic energy to the rotary ring of blades L while being deflected in the passages 9 between the rotary blades, it leaves the latter with an absolute exit speed c (exit angle a and the relative exit speed W2 (exit angle 5 in order to then enter into the adjoining unillustrated stationary ring of blades, or to discharge from the turbine casing, for example for the purpose of being conducted to an adjoining turbine component or to be extracted. From FIG. 3 it is clearly apparent that e+s=t-sine 5 if s indicates the width of the exit edge 8, which is to say where e is the opening ratio or the opening value.

In accordance with the above considerations, the speed diagram of FIG. 4, which is indicative of a reaction turbine having a degree of reaction on the order of r=0.5, can be readily understood. Because of the increase in the opening value c achieved with my invention, there is provided the illustrated increase in the mean entrance angle 04 and the increase in the mean exit angle [3 In the illustrated example the entrance angle a is equal to the exit angle p By way of the index 112 it is indicated that the angle is the mean value for the stationary and rotary blades and further that in the case of tapered or taperedtwisted blades it is also the mean value taken over the height of the blade. In order to better illustrate my invention there are also shown in dotted lines in FIG. 4 the angle u fl as being approximately equal to 26 corresponding to an opening value 6=0.45. The vectors of the absolute entrance and exit speeds of the stream, the relative entrance and exit speeds, and the rotary speed are also indicated at 0 c W1, W2, and the u respectively.

My invention has recognized the fact that a blade profile with large relative section modulus with a large opening value according to my invention can only be developed above a predetermined minimum Reyonolds number, namely 2X10 at the rated capacity. In this case the flow remains within the region of turbulence where the frictional drag of the primary stream at the boundary layer is sufiiciently great.

The illustrations of FIGS. 5a-5d show a constant normalized section modulus for the blade profile which is based on a dimensionless blade profile magnitude b which is to say these figures show tests of a profile family of the same constant resistance moment W/b Furthermore, these curves are derived with a blade profile where the quotient of ma divided by d' l o.7 with an opening value 6:0.3. Furthermore, corresponding practical factors have been taken into consideration, in that the gaps are proportional to the mean shaft diameter D the bearing distance L is proportional to D which is to say the root from the mean admission diameter and thus the critical speed n is approximately constant. Furthermore, it is assumed that the change of the active length of the rotor, which is to say that length which is provided with blades, equals the change in the distances between the bearings. Finally the Parsons number X is assumed to be constant.

As is shown in FIG. 5a, there takes place in a most surprising way with an increase in the opening angle 6 an improvement in the entire efficiency to such an extent that the'sum of all of the losses Zu h transferred to the ordinate in percentage, is smaller and has its minimum at approximately 5:0.6. This minimum can, depending upon the selected profile family, be displaced upwardly and downwardly to a small extent.

In FIG. 51) there are indicated the improvements in percentage of the pertinent individual losses depending upon the opening value 6. And thus there is the piston gap loss 11 indicated by the radial gap between the compensating pistons and easing of a predetermined amount of leakage flow. With high pressure turbines, as is known, it is necessary to provide such compensating pistons to equalize the axial thrust.

The blade gap loss a is derived from the fact that the radial gap between the blades and easing will result in a certain amount of leakage flow. The blade tip loss v is determined by the flow loss at the tips of the blades. The entrance flow value v, is determined by the following consideration: if the opening angle 6 is increased, then with constant blade losses there will be a decrease in the circumferential efiiciency, as already explained. On the other hand as result of the decrease in the deflection which then takes place in the passages between the blades there is a decrease in the blade loss. The influence of both of these changes on the efiiciency is indicated by the entrance flow value 11,. An important discovery of my invention resides in the fact that, as shown in the diagram, with increasing there is not only a decrease in the piston gap loss 11 and the gap loss 11 but also there is a decrease in the blade tip loss v so that the reduction in the individual efiiciency v, initially has no influence on the entire efiiciency and is more than compensated for, and in fact it is only at relatively large e-values that there is an actual reduction in the total efficiency. It is important that the blade tip loss v be proportional to the ratio of b to l, which is to say the axial blade width to the blade height or blade length. The behavior of these individual losses provides the curves shown in FIG. 5a for the total efficiency.

FIG. 5a shows that it is possible to reduce the blade height 75 or 1 because the increased opening value provides a larger passage section for the steam between the blades. At the same time it is possible to reduce the mean admission diameter 5. For the rotary blades, based upon a predetermined above-mentioned centrifugal force stressing at e=0.3, it is clear that the blade width b can be reduced, because the centrifugal force stresses and thus the ratio CF to a becomes smaller and for this reason, as shown in the diagram of FIG. 50, the permissible bending stresses (1 increases at the blade roots. Of course, because of the constant Parsons number it is necessary also to increase the numbers of stages, as shown by the curve Z in the diagram of FIG. 5c, but the extent of reduction of blade width is greater, and it is possible to reduce the active blade length of the rotor L as may be seen from FIG. 5d. Therefore, since not only the mean admission diameter but also the active length of the rotor are capable of being reduced, there is a highly significant saving in the cost of construction.

Referring now to FIG. 6, there is indicated, as mentioned above, the improvement in heat consumption which can be achieved in a high pressure, intermediate pressure, and low pressure casing of a 1000 MW-turbine, depending upon the weight diameter in the second casing DGIMD, indicated in mm., further depending upon the opening value or difierent sizes of the entrance angle sine a '=SiI1e {3 with a pair of different critical speeds as parameters. The further indicated pressure numbers P are defined by the dimensionless quotients 2 A is st/ 11 where A1, st represents the isentropic enthalpy diiterential of a given stage and a signifies the square of the circumferential speed. The pressure number I is, in the same way as the Parsons number X a characteristic magnitude of the turbine and a determining factor for the stage load. The product of X and ta is, as is known, equal to 8380 m kg/ From 6 it can be seen that with a constant critical speed and constant weight diameter of the first stage, which is to say with constant entrance fiow magnitudes, the operating safety, with increasing mean discharge flow angles 8 increase the gain in heat consumption. Since the distance of the individual curves for sine a =sine fi =0.3, 0.4, 0.5, 0.6 with increasing 6 becomes constantly smaller, it is clear that with e-values above 0.6 a saturation has been achieved which is to say in efiiciency has been reached where in the illustrated example it is at a maximum at values of 6 above 0.6. It the weight diameter is increased or the critical speed decreased, then, as is shown in FIG. 7, with a constant discharge flow angle the stage loading Will become reduced and the gain in heat consumption will be increased. In order to arrive at an optimum efficiency, it is recommended in accordance with my invention that the relative maximum profile thickness of the blades d /L be between 0.3 and 0.4, where ai is the diameter of the blade profile where the inner circle is situated at its thickest region and L is the profile chord length (the longest dimension of the blade section), as indicated in FIG. 1. In this way it is possible to achieve a highly as indicated in FIGS. 7 and 8, by using, instead of cylindrical-radial blades Pr a tapered, or better yet a taperedtwisted blade Pr As is shown in FIG. 8 the cylindricalradial blade Pr is characterized by the fact that the blade cross section IIII and the blade cross section II is constant throughout the entire height of the blade and the center of gravity thereof is situated on a radial line which extends perpendicularly across the axis of rotation. Thus, these sections of the blades shown at the left of FIG. 8 for the blade Pr are equal to each other. A tapered blade is characterized by the fact, as shown for the blade Pu in FIG. 8 that the cross section of the blade diminishes gradually toward the tip thereof, so that the cross section in the region of the tip of the blade is smaller as indicated by the same sections taken for the right blade of FIG. 8 as compared to those taken for the left blade of FIG. 8. A tapered-twisted blade construction provides still further advantageous features in that the individual sections are turned one relative to the other as indicated by the arrow at the upper right portion of FIG. 8 which shows a turning of one blade section relative to the other to provide the twisted blade structure. As a result of this piston blade structure, as compared to the cylindricalradial structure of FIG. 8, Where the blade has a constant cross section throughout its length, it is possible to achieve the improved heat consumption indicated in FIG. 7 where with the same type of illustration as shown in FIG. 6 D MD indicates again the weight diameter in the second or intermediate pressure casing and 1 indicates the pressure number. It is seen, therefore, that the blade structure Pr achieves a highly significant improvement in heat consumption as compared to the blade structure Pr The critical rotary speed again is maintained constant. In order to indicate the changes in the gap-piston and discharge losses in a clearer manner, the improvement of the flow losses resulting from the use of taperedtwisted blades, which results in the greatest fraction of the total improvement, is not shown in FIG. 7. The actual improvement of heat consumption is therefore greater than illustrated. A tapered or even better a taperedtwisted blade structure Pr is thus to be used with advantage in the structure of my invention. Even if such tapered-twisted blades for reasons of strength are preferably used with advantage for rotary blades of relatively great centrifugal force stresses, nevertheless it is preferred to use such blade structures also for rotary blades subjected to lesser extents of centrifugal force stresses as well as for stationary blades because of the improvement in efficiency resulting from the use of such blades.

My invention is not to be considered as limited to use with blades for turbines where steam is the operating fluid under pressure, since it is also capable of being used with turbines where instead of steam the operating fluid under pressure is, for example, ammonia (NH or Freon, for example Freon 21 (CHClF Thus, the range of uses of my invention are only limited by such factors as for a given rated capacity of a turbine stage the increase in the specific volume of each stage from one to the next is at a maximum 15% and the Reynolds number is equal to or greater than 2x10 I claim:

1. In a turbine, at least one stationary ring of blades and a rotor carrying a rotary ring of blades for rotary movement relative to the stationary ring of blades, said stationary and rotary rings forming at leasrt one of a series of turbine stages arranged one after the other in the direction in which the pressure of the turbine-driving fluid drops while the turbine-driving fluid expands during rotary driving of the rotor, said blades having at their mean diameter at a region of a given operating factor and a Reynolds number Re=w -b/v*;2 10 at the rated capacity of the turbine stage, an opening value e=e/t at least equal to 0.45 as the mean value for the stationary and rotary blades, where e=the smallest distance between adjoining blades,

1 1 t=the distance between corresponding points of a pair of adjoining blades, w =the relative exit speed, b=the blade Width, and v*=the kinematic viscosity.

2. The combination of claim 1 and wherein said given factor at said region is a pressure which is at least equal to 2 atmospheres absolute.

3. The combination of claim 1 and wherein said given factor at said region is an increase of not more than 15% in the specific volume of each stage.

4. The combination of claim 1 and wherein the ratio 6:6/[ is between 0.50 and 0.75.

5. The combination of claim 4 and wherein said ratio is between 0.60 and 0.70.

6. The combination of claim 1 and wherein a highpressure region of the turbine is designed to operate at a mean pressure on the order 90 atmospheres absolute.

7. The combination of claim 1 and wherein an intermediate mean region of the turbine is designed to operate at a mean pressure on the order of 8 atmospheres absolute.

8. The combination of claim 1 and wherein the extent of reaction of the turbine is between 25 and 65%.

9. The combination of claim 8 and wherein the extent of reaction is on the order of 50%.

10. The combination of claim 1 and wherein the turbine is an axial turbine.

11. The combination of claim 1 and wherein the relative maximum profile thicknessc of the blades d /L is between 0.3 and 0.4, wherein d is the diameter of blade profile at the inner circle arranged at its thickest region and L is the longest chord length of the blade profile.

12. The combination of claim 11 and wherein at least the rotary blades are tapered.

13. The combination of claim 12 and wherein at least said rotary blades are also twisted.

14. The combination of claim 1 and wherein a fluid under pressure which coact with said blades is selected from the group consisting of steam, ammonia (NH and Freon 21 (CI-IClF References Cited UNITED STATES PATENTS 1,539,395 5/1925 Losel 25377 1,777,098 9/ 1930 Lysholm.

FOREIGN PATENTS 1,018,407 10/ 1952 France.

1,264,388 5/1961 France.

EVERETTE A. POWELL, 111., Primary Examiner 3 UNITED STATES PATENT OFFICE CERTIFICATE OF CORRECTION Patent No. 3, l-75,1O8 Dated October 213 1969 Inventor(s) Waldemar Zickuhr It is certified that error appears in the above-identified patent and that said Letters Patent are hereby corrected as shown below:

I In the heading; to the printed specification line 3,

"Walemar" should read. Waldemar SIGNED AND SEALED JUN 9 197 (SEAL) Attest:

Edward member WILLIAM 2. mm, m. Attesting Officer Commissioner of Patents