ISSN 00406015, Thermal Engineering, 2010, Vol. 57, No. 12, pp. 1042–1051. © Pleiades Publishing, Inc., 2010. Original Russian Text © G.M. Morgunov, 2010, published in Teploenergetika.

Improvement of the Main Pump Equipment Used in Large Thermal Power Installations G. M. Morgunov Moscow Power Engineering Institute, ul. Krasnokazarmennaya 14, Moscow, 111250 Russia Abstract—The establishment and development history of the Moscow Power Engineering Institute’s Department of Hydromechanics and Hydraulic Machines, and the educational and scientific work presently conducted at it are briefly described. The research and design work aimed at developing the optimal construc tion of the flow path and blade systems for the stages of a feedwater pump with a useful capacity of 25 MW with the traditional onesided admission of working fluid for power units designed to operate at supercritical steam conditions is discussed. A new alternative design solution for the cartridge of a superpowerful feedwa ter pump with the twosided admission of working fluid is presented. The predicted possibility of achieving more reliable and energyefficient performance of a feedwater pump equipped with such a cartridge is sub stantiated. DOI: 10.1134/S0040601510120086

In connection with a jubilee issue of this number, the author considers it appropriate to present, in the form of a brief essay, the main milestones of the peda gogical and scientific work conducted at the Depart ment of Hydromechanics and Hydraulic Machines (HHM) with permissible abridgements. A BRIEF ESTABLISHMENT HISTORY OF THE DEPARTMENT OF HYDROMECHANICS AND HYDRAULIC MACHINES AND ITS SCIENTIFIC AND PEDAGOGICAL WORK The Department of HHM is a graduateissuing department educating students on the specialty 121100 “Hydraulic Machines, Hydraulic Drives, and Hydropneumatic Automatic Devices,” namely, grad uated engineers specializing in the discipline “Hydraulic, Vacuum, and Compressor Engineering,” as well as bachelors and masters specializing in the dis cipline “Construction of Power Machinery and Equipment.” Engineers are educated for two special ties: (1) engineering design and investigations in con struction of hydraulic machinery and (2) erection, adjustment, and operation of hydraulic equipment. The following two important circumstances are wor thy of noting. (i) The entire educational process conducted at the department is based on an approach oriented at sys tematic and creative activities taking into account the present social and economic situation in the country and placing focus on the humanitarian and environ mental components of education. (ii) It is difficult to draw up a more or less complete list of modern production facilities at which hydraulic machines, systems, and hydropneumatic automatic

control devices are used; thermal power facilities would surely occupy a prominent place in such a list. Experience shows that to a certain degree owing to the abovementioned specific features of education, as well as due to the fact that specialists in this field are in wide demand in the market of highskilled staff, grad uates from the department have an increased level of competitiveness and adapt comparatively easily to constantly changing conditions of market relations in the country. Demand for graduates from the depart ment in the industry and at scientific organizations is stably high, and job finding matters are quite solvable for them. Historically, the HHM Department was estab lished in 1982 by merging the departments of hydrau lics and hydraulic machines, which were founded in 1945–1946 within the hydraulic power engineering faculty that had been set up at the Moscow Power Engineering Institute (MEI). The Professor, Doct. Techn. Sci. V.S. Kvyatkovskii, was the first head of the department of hydraulic machines (up to 1974) and Professor, Doct. Techn. Sci. S.V. Izbash was the first head of the hydraulics department (up to 1972). These prominent figures headed the work on estab lishing and then developing a system of tutorial and methodical aids for supporting the education pro cesses, as well as laboratory facilities with the neces sary materials, equipment, and instruments for effi cient performance of education and scientific works. This work was carried out at the corresponding depart ments from the time of their establishment until approximately the first half of the 1960s. In doing this difficult and multisided work, the heads of the depart ments found support from the remarkable group of firstgeneration scientists and teachers. Here and

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IMPROVEMENT OF THE MAIN PUMP EQUIPMENT USED

henceforth, the generations of specialists are related to different intervals of time in a rather conditional man ner, and the participants of education or scientific schools are given mainly in alphabetic order without indication of their scientific titles or degrees. The following people constituted the backbone group of specialists at the department of hydraulics: B.T. Emtsev, I.V. Lebedev, N.M. Leleeva, P.M. Slis skii, S.M. Slisskii, A.I. Smolyak, and K.Yu. Khaldre. The following was done owing to their efforts: (i) The education course “Hydraulics” was estab lished and developed (Professor Izbash), and evolu tionarily transformed (Professor Emtsev) into the dis cipline “Fluid Mechanics” including its support with laboratory works that were at an advanced level at those times; this course was given at a few faculties of MEI. (ii) Scientific teams on the problems of hydraulic engineering were established, and their fruitful work was organized (Emtsev, Izbash, and S.M. Slisskii). Generalizing publications, predominantly in the form of monographs, issued in that and subsequent period of times, including publications of a similar level written at the department of hydraulic machines, became handbooks of engineers on hydromechanics specializing in this field [1]. As regards the department of hydraulic machines, the following people constituted the core of its scien tific and pedagogical staff: G. V. Viktorov, B. E Gleze rov, A.N. Mashin, A.P. Sokolov, and M.M. Ora khelashvili, and then O.A. Verbitskaya, B.I. Shvarts burd, and I.G. Yan’shina. The following achievements can be related to the greatest results of their joint work: (i) Educational disciplines on hydraulic machines and hydraulic drives were established at an advanced level at those times. (ii) Unique (at those times) power cavitation rigs for testing model hydraulic turbines, blade pumps, and first education setups for studying positivedisplace ment hydraulic machines and drives were constructed. (iii) Scientific avenues were shaped, and fruitful theoretical and experimental activities were com menced in the fields of investigation and design of hydraulic turbines and pumps, predominantly diago nal adjustableblade ones. (iv) The topic of positivedisplacement hydraulic machines and hydraulic drives, which was a new scien tific and education course, began to be developed from the late 1950s due to fruitful efforts taken by B.E. Glezerov and G.V. Viktorov, and then V.A. Lesh chenko. The first doctoral (I.V. Lebedev and B.T. Emtsev) and candidate’s dissertations were presented at both the departments. The period of time from approxi mately the mid1960s to the early1980s was the time in which the departments of hydraulics and hydraulic THERMAL ENGINEERING

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machines were in their flourish. The education process was constantly improved, the fundamental disciplines of engineer training were deepened, new special courses were introduced, and qualitative improve ments were made in the laboratory facilities at that time. Specialists from research institutes and industry were attracted to pedagogical and scientific work (O.F. Nikitin, Yu.A. Petrov, V.I. Razintsev, and others). Special mention should be made about the brilliant work of Professor M.M. Orakhelashvili when he was the dean of the Faculty for Construction of Power Machinery and Equipment, which incorporated the department of hydraulic machines in its composition in 1961. In 2010, the 100th anniversary from the birth day of this unforgettable outstanding person, a scien tist, teacher, and prominent man, will be celebrated. At the time Orakhelashvili headed the faculty the fol lowing scientific schools emerged and conducted fruitful work at the department of hydraulics: (i) the school of applied hydromechanics (Emtsev as a scientific leader, A.I. Smolyak, then E.I. Pyatigor skaya, and then younger researchers V.V. Bernev, A.A. Karpukhin, and others); and (ii) the school of jet elements and hydropneumatic automatic control devices (I.V. Lebedev as a scientific leader, V.S. Levin, S.M. Treskunov, V.S. Yakovenko, and then A.I. Davydov, V.L. Ostrovskii, and others). At that time, the following fields were developed at the department of hydraulic machines: (i) calculation, designing, and experimental study ing of diagonal adjustableblade hydraulic machines, namely, turbines, pumps, and reversible hydraulic machines (V.S. Kvyatkovskii as a scientific leader, G.V. Viktorov, B.E. Glezerov, A.N. Mashin, A.P. Sokolov, then O.A. Verbitskaya, I.G. Yan’shina, then I.G. Belash, S.N. Pankratov, D.Kh. Tsakiris, and then B.M. Ora khelashvili, M.I. Kosenkova, and others); (ii) development of theory and working out numer ical methods for solving direct and inverse 2D nonpla nar hydrodynamic problems for the working members of blade hydraulic machines followed by statement and calculation of quasi 3D problems of blade cas cades (G.V. Viktorov as a scientific leader, I.V. Vychk ova, A.F. Makovetskii, G.M. Morgunov, and others); subsequently, the author of this paper led this field working jointly with his colleagues A.V. Volkov, V.M. Gorban’, V.V. Frolov, S.N. Shepelev, and others; (iii) investigation, designing, and static and dynamic simulation of positivedisplacement hydrau lic machines, throttle and volume hydraulic control systems (in the 1960s led by V.A. Leshchenko, and in subsequent years by V.I. Golubev and Yu.Yu. Zuev with their followers M.M. Zaverskii, A.M. Popov, E.N. Sorokin, and others); and (iv) modern technology for manufacturing hydrau lic machines (B.I. Shvartsburd as a scientific leader, V.I. Golubev, A.N. Dunaev, and M.M. Khachaturov).

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In the 1960s and in the subsequent years, a few doctoral (G.V. Viktorov, G.M. Morgunov, and A.V. Volkov) and a few tens of candidate’s dissertations were presented. The departments of hydraulics were headed by Professors, Doctors of Technical Sciences B.T. Emt sev (1972–1982) and G.V. Viktorov (1974–1982). It is easy to notice that more and more new names appear in the lists of people participating in scientific and education fields of work. These are new promising specialists, many of which subsequently determined the main staff of separate departments, and since 1982, the united department of hydraulics and hydraulic machines. It should be pointed out running a few steps forward that in 2009, in view of outstanding scientific, pedagogical, and organizational services made by the first head of the department of hydraulic machines, Professor, Doctor of Technical Sciences, twice the laureate of the USSR State Award Vladimir Stanislavovich Kvyatkovskii, the department of hydrau lics and hydraulic machines was named after him. After the two departments had been united, the HHM department was headed by Professors B.T. Emtsev (from 1982 to 1987), G.M. Morgunov (from 1987 to 1995), V.I. Golubev (from 1995 to 2004), and A.M. Gribkov (from 2004 until nowadays). During a considerable period of time (since 1982 and until nowadays), the HHM department carried out transformations of the education process and sci entific work common for the entire system of higher education in the country connected with scientific technical progress, entering into market relations and into the era of computer and information technologies unwitnessed in the scales and quality of their develop ment. Cardinal changes were made in the educational plans and teaching methodology in connection with making a shift to a multilevel system of training, which was mainly determined by the conditions of its fitting into educational structures and subsequent highly adaptive professional activity of graduates adopted in some advanced western countries. Nonetheless, we cannot but point out the fundamental novelty of the statement and implementation of systematically cre ative education procedures developed at the department, which were implemented out in the fullest form in 1987– 1993 (Yu.Yu. Zuev, V. S. Levin, and G. M. Morgunov) jointly with the Center of Engineering Design (Profes sor V.F. Vzyatyshev) that actively worked at that time. Since it is not possible to describe this complex educa tional experiment, which nonetheless has positively proven itself on the whole, in a more or less full detail in this publication, interested readers can find more information on this topic in the generalizing mono graphs [2, 3]. The scientific work that was conducted in those years can be characterized by the following prominent achievements.

In the scientific area of fluid mechanics and theory of hydraulic machines, mathematical models were developed for describing spatial unsteady (in the Rey nolds sense) turbulent flows, central to which is the use of integral equations and representations of the multi dimensional field theory (G.M. Morgunov). In the scientific area of blade machines: (i) A range of diagonal adjustableblade turbines was developed (V.S. Kvyatkovskii, and also I.G. Belash, A.P. Sokolov, and D.Kh. Tsakiris). (ii) An innovation group for constructing the main turbine equipment for small hydraulic power stations (HPSs) was established and is actively working (B.M. Orakhelashvili). (iii) Innovative design solutions were obtained with substantiation of competitive properties for new lay outs of HPSs, the design of radialaxial hydraulic tur bines, turbines for directly using the kinetic energy of fluid media, and multirow single and multistage blade machines (G.M. Morgunov). In the scientific area of positivedisplacement hydraulic machines, hydraulic drives, and hydropneu matic automatic control devices: (i) A new hydraulically driven system for wind power installations has been proposed (V.I. Golubev and I.A. Zyubin). (ii) The advanced concept of a fully electrified plant equipped with hydraulic energy generation and consumption devices has been developed. (iii) A standard series of such systems and pilot models have been developed (Yu.Yu. Zuev). A more detailed description of the history of devel opment and subsequent scientific and pedagogical work of the HHM department, which will turn 63 in this year, with an extended list of papers published by its workers is given in [1]. Below, the concrete topic of this publication is considered. INVESTIGATIONS AND PROJECT DEVELOPMENTS FOR THE SUBSYSTEM OF FEEDWATER PUMPS Feedwater pumps used in powergenerating instal lations make up the most energyintensive and impor tant group of pump systems used in the main process cycles. It should be noted that the stable tendency toward an increase in the capacity of single power units entails the corresponding increase in the capacity of feedwater pumps. The list of the most important requirements imposed on these hydraulic machines includes, apart from ensuring high energy efficiency, the need to ensure the highest possible reliability, max imal service life, and technical availability. On one hand, an increase in the efficiency of 40MW feedwa ter pumps by only 1% makes it possible to produce an additional 3.5 × 106 kW h of electric energy per annum. On the other hand, forced shutdown of a THERMAL ENGINEERING

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2

k9

3

56

1

56 ∅565

38

∅430

k9

∅280

k9

∅270

∅420

4 ∅180

1000MW power unit for 3.5 h to carry out shortterm repairs to its feedwater pumps entails an equivalent loss in the production of electricity. In particular, according to the statistical data of the North American Council for Reliability, faults or failures of feedwater pumps turned to be at the third place by number among the factors that caused outages of large thermal power stations, which in many respects predetermined the cost of underproduced electric energy exceeding $400 million [4]. Thus, the most important objective that has to be fulfilled in designing newgeneration feedwa ter pumps is to achieve the highest hydraulic power, efficiency, and reliability while keeping an optimal ratio between them [5]. In accordance with this objective, an attempt was made to carry out research and design work on devel oping the active part of the cartridge for a feedwater pump of the traditional structural makeup, i.e., with onesided supply of working fluid (feedwater), as wells as with stages including a centrifugal runner and two guide vanes: a direct guide vane (DGV) and a reverse guide vane (RGV) equipped with a system for taking the axial force by a relief disk (RD). The operating parameters were determined proceeding from the actual intervals of their values and were found to be as follows:

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Fig. 1. Studied stage of the feedwater pump in the meridian projection. (1) Runner, (2) DGV, (3) RGV, and (4) shaft.

Developed hydraulic (i.e., useful) power Nd, No less than 27 MW Delivery Q, m3/s 0.765 Number of stages m 5 Stage head Hs, m 740 Angular rotation frequency ω, 1/s 624 Working fluid temperature at the inlet t1, °C 160 Working fluid pressure at the inlet p1, MPa 2.5 Stage hydraulic efficiency ηh.s No less than 0.86

These objectives and the tasks stemming from them generated the need to set up a design and calculation computer experiment, which was carried out in an interactive mode with using the wellknown 3D method for describing flows in the blade systems of turbine machines in performing hydrodynamic calcu lations [6].

The research and design work pursues the following objectives: (i) determining the geometrical parameters and shapes of the stage runners, DGVs, and RGVs, as well as of the system of relief disks, the use of which makes it possible to obtain the desirable integral indicators during operation in the nominal mode and with devi ations from it with respect to delivery by approxi mately ±50% with ensuring the condition for cavita tionfree operation of the first stage; (ii) carrying out hydrodynamic calculations and studying the distributions of field functions with estab lishing methods for optimizing them in terms of improving power performance indicators and reduc ing vibration characteristics of the feedwater pump’s rotor; (iii) possible adoption of alternative competitive solutions for the feedwater pump and its prebooster devices.

The fact that the flow path used in the stages of the feedwater pump produced by Sulzer [4] was selected from its determining geometrical relations as a tenta tive analog in designing the feedwater pump cartridge can be regarded as one of the most essential results obtained from these calculations. As an example, the dashed lines in Fig. 1 show the axially symmetrical surfaces k9. The distributions of the velocity modulus (relative velocity w for the runner and absolute velocity v for the DGV and RGV), the pressure coefficient p, and the friction stresses τfr on the blades in the rated mode of operation are shown in Fig. 2 (see [6]). The table summarizes the determining parameters of the feedwater pump’s working members in the studied range of delivery from 0.35 to 1.20 m3/s, and Fig. 3 shows the predicted integral performance characteris tics of the stage. The characters Γ1 and Γ0 in the table denote the flow circulation at the inlet to the corre sponding blade system (cascade) and the circulation created by this cascade, Hth.st and Hst are the theoreti

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MORGUNOV ω

W2

(a) 0.5 × 10–2w, m/s; p; 10–2τw, kN/m2 0.6 2 1

0.4 0.2

–1.0

–0.5

3

0 –0.2

0.5

s

W1 (b) 22°

1.0 10–2v, m/s; p; 10–2τw, kN/m2 0.8 0.6

∅ 570 ∅4 30

10°

2

0.4

1

0.2 3 –1.0

–0.5

0.5

0

s

(c) 10–2v, m/s; p; 2.0 10–2τw, kN/m2

20°

1

1.5

∅ 5 56

∅2 80

94 °

1.0 p 1.00 0.5 0.99

3

2

0.98 –1.0

–0.5

0

0.5

s

Fig. 2. Distributions of field functions in k9 sections. The “+” and “–“signs denote the working and rear sides of blades (or pro files in the k9 sections), s is the dimensionless meridian projection of the corresponding blades in the k9 sections (taking into account the near wake). (a) is the axonometric image of the runner blade, (b) is a fragment of the cylindrical blade cascade of the direct guide vane DGV, and (c) is a fragment of the cylindrical blade cascade of the reverse guide vane RGV. (1) w or v, (2) p, and (3) τw.

cal head and the expected head obtained from predic tion assessments, Nst is the hydraulic power of the stage, kc is the cavitation parameter determining the necessary positive suction head at the inlet to the first

stage, hh is the hydraulic losses in the cascade, and ηh.st is the hydraulic efficiency of the stage. It can be seen in the lefthand parts of Figs. 2a–2c that the blades of the runner have a spatial shape, and THERMAL ENGINEERING

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Determining parameters for the feedwater pump’s working members in the studied range of delivery Q, m3/s

Γ1, m2/s

Γ0, m2/s

Hth.st, m

hc, m

hh, %

ηh.st, %

Hst, m

Nst, MW

0.350

0.0

92.3

935

66.4

6.5

83.4

780

2.68

DGV

92.3

–73.9

–

6.8

RGV

18.4

–15.7

–

3.3

0.0

83.0

118.6

5.7

86.0

723

5.43

DGV

83.0

–53.4

–

6.3

RGV

29.6

–26.0

–

2.0 84.6

621

7.31

Working member Runner

Runner

Runner

0.765

1.200

841

0.0

72.4

176.9

6.5

DGV

72.4

–32.6

734

–

6.7

RGV

39.8

–35.7

–

2.2

the blades of the DGV and RGV have a cylindrical shape. To reduce the possibility of exciting highfre quency vibration of the rotor caused by interaction between the wakes downstream of the runner blade exit edges and the DGV and RGV inlet edges, the number of blades in these members is taken to be as follows according to the recommendations of KSB [7]: 7 in the runner, 9 in the DGV, and 12 in the RGV. An analysis of the distributions of field functions, specifically, the diagrams shown in Fig. 2, allows the following conclusions to be drawn. The velocities w (see the diagrams shown in Fig. 2a) on the rear side of the runner blades are rather high (their average value is around 90 m/s) and increase monotonically towards the exit. The values of w on the working surface increase rapidly almost immediately behind the inlet edge. Such dependences correlate with the levels and pattern of change in the stresses τw. The pressure coef ficient p, increases, as should be expected, toward the exit on both sides of the blade thus reflecting a diver gent pattern of flow in the runner. A negative factor that should be pointed out is that the considered func tions have essential outbursts as working fluid flows over the inlet edge, thus causing some degradation in the power performance indicators of the stages (the socalled shock losses increase) and in the anticavita tion properties of the firststage runner (see the zone of negative p on the diagrams shown in Fig. 2a). The length of the interblade channel, the diver gence ratio (see the lefthand part of Fig. 2b), the velocities v, and the stresses τw on the rear side of the DGV blade (see the righthand part of Fig. 2b) all have significant values. The velocities v on the RGV blade are comparatively low, and outside of the exit edge vicinity they are approximately equal to the meridian component of absolute velocities. The friction stresses τw are also considerably smaller than those in the run ner (see the righthand part of Fig. 2c). Another thing that attracts attention is that the pressure increases permanently from the runner toward the RGV, with p THERMAL ENGINEERING

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reaching a value equal to around unity at the inlet to the next stage. The value by which p differs from unity is determined only by the kinetic energy of meridian flow. Thus, it follows from the results of the numerical experiment that the designed stage is predictively in line with the initial nominal parameters with respect to power performance indicators (Nd and ηh.st) and anti cavitation characteristics (the cavitation margin at t = 160°C and, hence, at the head of saturated vapor Hs.v ≈ 63 m) will be equal to 15 m. The head–flowrate characteristic Hst = f(Q) has a monotonically decreas ing pattern (see Fig. 3). However, the condition ηh.st ≥ 0.86 is satisfied only in the form of a weak inequality. The presented distributions of field functions point to the possibility of further improvement in the blade streamlining pattern. In particular, it is advisable to correct the cascade density and the inlet parts of work ing member blades to reduce the peak level of veloci ties. This is especially the case for the DGV (see the table). Hst, m hc, m ηth.st

Nst, MW

800 600 400 200 0

Hst 7

90 ηth.st

5

70

Nst

50 30 0.2

hc

3 1

0.4

0.6

0.8 1.0 Q, m3/s

Fig. 3. Predicted working characteristics of the stage in the vicinity of the rated mode of operation.

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MORGUNOV From the condensation system 1 2 ρ < ρa

3

ρ > ρw

By A

From 3

From 3

A By B To the steam generator

From 4

4

5

B

a1

b1

c1

c2

b2

a2

Fig. 4. Circuit fragment of the power installation with twosided hydraulic pump lines. (1) Deaerator, (2) backing receiver, (3) flow splitter, (4) separating booster pump, and (5) DFWP.

Nonetheless, as is shown quite convincingly in [7], considering the experience gained around the world in construction of competitive feedwater pumps, it is unlikely that the overall efficiency could be improved by more than 2.5% solely by improving the hydrody namic characteristics of their flow paths. This circum stance, as well as the wellknown facts that additional losses of mechanical energy and frequent failures of feedwater pumps during their operation are connected with the device for equalizing axial loads on the rotor [4, 5, 7, 8] prompted specialists to design and carry out an indepth study of the unbalanced axial forces from the runner that have not been equalized by means of a relief disk at the nominal parameters of the pump mentioned above [9]. The final results obtained from that development have shown that the axial force on the feedwater pump’s rotor is equal to 0.9 MN and is equalized by a relief disk with the endface slit’s outer and inner diameters equal to 0.4 and 0.3 m with the endface gap equal to 137 μm. The flowrate through the relief disk system under these conditions was equal to 4% of the pump’s total delivery with the corre sponding drop in the useful power. There are data (see, for example, [7]) indicating that the axial force arising in feedwater pumps of a higher capacity (more than 40 MW) exceeds 1.0 MN, and the working range of the end gap fits into the interval 0–50 μm; i.e., the relief disk has an extremely steep static load characteristic. Hence, the probability of its coming in direct contact with the stator disk surface increases during operation. In view of the above, a prospective attempt was taken to develop the design of a feedwater pump for superhigh parameters with twosided admission of working fluid and equalizing the residual axial force

only by means of a doubleaction axial bearing. This pump will be denoted by the conditional abbreviation DFWP. The data on the calculated operating condi tions of the DFWP are as follows: Nd = 45 MW, Q = 1.15 m3/s, m = 10 (by five on each side), Hst = 800 m (with the runner diameter D2 = 0.43 m), ω = 624 s–1, t1 = 190°C, p1 = 3.25 MPa, and ηh.st ≥ 0.88. The system supporting normal operation of the DFWP is schematically shown in Fig. 4. On passing treatment in the deaerator, feedwater is fed to a back ing receiver that serves to create thermodynamic con ditions required for cavitationfree operation of the subsequent devices. Thus, in the representations of equilibrium thermodynamics, when the temperature in the receiver drops adiabatically from 200 to 190°C, the ratio of pressure at its outlet to the pressure of sat urated vapor will increase by approximately a factor of 1.19. After that, the working fluid is directed to a spe cial device (flow splitter) that serves to divide the flow, as far as possible, into two equal parts. The flow splitter can be nonadjustable (similar to the design solutions traditionally used in Dtype centrifugal pumps) or can operate automatically according to an openloop or closedloop control circuit (the latter version is the most preferred one). After that, the split flow is sup plied to the inlets of the hydraulic machine’s separat ing booster pump running with the rotation frequen cies equal to half that of the DFWP. As follows from its name, the separating booster pump has a dual pur pose. First, it serves to ensure the required anticavita tion properties of the DFWP, i.e., the pressure p1 at its inlet. And second, it serves to mechanically clean the working fluid in addition to the thermal and chemical water treatment methods used in the preceding appa THERMAL ENGINEERING

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1049

(a) 11

12

10

1

3a 3b

3c

A

4

5

d

D1

D2

D0

2

6

7 8 (b) A–A(turned through 90°) P

9

12 1

N

Evolvent of the D0 circle 12°

2

15 °

0

M

5° 10°

L D

D

0

30°

° 30

O

K

Fig. 5. Fragment of the cartridge used in the DFWP feedwater pump with twosided supply of working fluid. (a) Section; (b) sec tion by the suction cavity; (1) inlet cavity; (2) guide cascade; (3a) additional radial bearing; (3b) system for cooling and lubricating (3a) with water; (3c) air chamber; (4) thrust ring; (5), (6), and (7) runner, DGV, and RGV of the regular stage, (8) oppositely doubled runner of the central stage, (9) shaft, (10) and (11) inner and outer RGVs of the doubled central stage, and (12) fragment of the DFWP outer shell.

ratuses from components with densities higher or lower than the feedwater density ρfw, including coagu lants or products of their decomposition when separa tion is carried out with anomalous density at the desired values of pH. The view A shown in Fig. 4 con ditionally marks out the following sections: a1 and a2 correspond to a lowhead pump stage with a high suc tion capacity (a truncated screw), b1 and b2 correspond to a separation drum with end channels for evacuating the abovementioned components, and c1 and c2 cor THERMAL ENGINEERING

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respond to a highhead pump that creates the pressure 1

p1 at the DFWP inlets (see, for example, [10]).

On leaving the hydraulic machine, working fluid is supplied to the DFWP suction sockets with a flowrate approximately equal to Q/2. Figure 5a shows the frag ment of the sectional view of its cartridge, which 1 A more detailed discussion of conceptual issues and all the more

structural and parametric solutions for devices 2–4 (see Fig. 4) is beyond the scope of this publication.

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explains the layout of stages in the active part, which are arranged symmetrically with respect to the vertical working fluid removal axis, and Fig. 5b shows section A–A over the lefthand inlet channel. The places in which working fluid is supplied to the cartridge and removed from it are marked in Fig. 5a by vertical arrows. Below, it is assumed that the purpose, design and technical makeup, and operation principle of the main assemblies and elements of the pump’s active part (see Fig. 5) are known. The shift for using twosided supply of working fluid with the same rotation frequency, head, and resulting delivery leads in an obvious way to obtaining better anticavitation properties of the pump and (theoretically) higher hydraulic efficiency due to lower meridian velocities. For achieving high power performance indicators and cavitation properties of the DFWP, it is supposed, in particular, to use the experience that has already been gained with designing and studying the characteristics of feedwater pump stages having the traditional layout. In addition, it is necessary to keep the stage rotation speed within the recommended range of specific reduced rotation fre quency equal to 27–38 [7]. With the calculated parameters of the DFWP, this frequency is equal to 30.2. To obtain flow at the inlets to the firststage run ners that has a uniform pattern and is maximally close to an axially symmetrical circulationfree stream and simultaneously reduce the axial size of the inlet cavi ties, their inner surface at the suction diameter D0 = D1 is made as an evolvent of the circle of this diameter (see Fig. 5b). Such a spiral has the following remark able property: the LM segment has a length equal to the length of the LK arc. Clearly, ON = πD0/2. The point P is the head of the evolvent’s inlet edge. Figure 5b also conditionally shows the diameter D0 of the inlet socket on the DFWP outer shell. However, the use of the system with twosided sup ply of working fluid involves serious difficulties. The main of them is that the DFWP cartridge’s active part becomes almost a factor of 2 larger. This results in that the rotor partly loses its dynamic strength and becomes less stable to vibration. To make the shaft stiffer and, consequently, to reduce its natural sag under the effect of radial loads, the hubtotip ratio in the active zone (see Fig. 5a) has been increased from 0.38 to 0.46. For making it possible to develop the required head without increasing the diameter D2, the runner is made with a biplane blade cascade. The inlet edges of the additional deflector periodic system of runner blades are shown in Fig. 5a by transverse dashed lines. As is known, the use of such a blade sys tem makes it possible to increase the runner head by 10–13% (see, for example, [11]). To keep the vibration characteristics at their level in feedwater pumps of the traditional design, an interme diate radial support is added on each side (item 3a in

Fig. 5a) between the end seal (not shown in Fig. 5) and the shell of guide vane cascade 2. As a result, the dis tances between the supports in the DFWP can be made even shorter as compared with those in the feed water pump with a onesided supply of working fluid. It is supposed to install an elastic metal–plastic fric tion bearing [12] with organizing an air “barrier” in the form of annular cavity 3c [4]; the pump is cooled and lubricated by feedwater from a separate lowtem perature reservoir. It has been shown on the basis of calculated theo retical and experimental data reported in [12] and other publications of this scientific school that water lubricated friction bearings with the elastic effect in selfaligning segments have a high bearing capacity (3.5 MPa at a sliding velocity of 62.8 m/s), good dura bility of operation, smaller friction losses (by a factor of 3–4), and many other positive features as compared with the traditional sliding supports equipped with rigid, e.g., whitemetal bushings and with oil lubrica tion. If these properties of elastic bearings are con firmed in many experiments, a conclusion can be drawn about the advisability of using them as axial and end radial supports for the DFWP rotor. In the opinion of the author of this paper, by apply ing the totality of the adopted solutions it will be pos sible to construct feedwater pumps outperforming the existing ones in installed capacity, power performance and cavitation characteristics, and reliability. It should be pointed out that only one method for secondary placement of two sections of stages was considered in this paper, namely, when they operate in parallel and with the flow splitter installed upstream of the booster hydraulic machine. Other layout versions are possible with a different degree of preference of design solu tions, e.g., with sequential operation of both the sec tions. The author thanks students M.M. Piskunov and V.V. Klimova (Zharkova) who took efficient participa tion at certain stages of calculations and design work. The author also hopes that the described design and research work will raise interest among specialists working in the field of pump construction for power engineering applications. REFERENCES 1. V. I. Golubev, B. T. Emtsev, Yu. Yu. Zuev, and G. M. Morgunov, “The Scientific and Pedagogical School of the Department of Hydromechanics and Hydraulic Machines,” Vestnik MEI, No. 5, 35–39 (2003). 2. G. M. Morgunov, Sociosynenergy and Education (MEI, Moscow, 2005) [in Russian]. 3. Yu. Yu. Zuev, Principles of Constructing Competitive Equipment and Working Out Efficient Solutions (MEI, Moscow, 2006) [in Russian]. THERMAL ENGINEERING

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IMPROVEMENT OF THE MAIN PUMP EQUIPMENT USED 4. P. Bushziper, “The Design Concept of Feedwater Pumps Produced by SULZER,” Vestnik YuUGTU, No. 1 (41), 65–72 (2005). 5. D. Khelmann, “Matters of Working Out the Optimal Designs of Large Centrifugal Pumps for Thermal Power Stations,” Vestnik YuUGTY, No. 1 (41), 25–31 (2005). 6. G. M. Morgunov, “Calculation of Currents Flowing Over Spatial Blade System without Separation Taking Viscosity into Account,” Izv. Akad. Nauk USSR, Ener getika i Transport, No. 1, 117–126 (1985). 7. Yu. Shil’, “Tendencies in the Development of Feedwa ter Pumps,” Vestnik YuUGTU, No. 1 (41), 32–36 (2005). 8. Blade Pumps: a Handbook, Ed. by V. A. Zimnitskii and V. A. Umov (Mashinostroenie, Leningrad, 1986) [in Russian]. 9. V. V. Klimova and G. M. Morgunov, “Static Calcula tion of Axial Forces in the “Active Part–Balancing Ring” System of a Centrifugal Pump,” in Proceedings of

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International ScientificTechnical Conference “Effi ciency and Environmental Characteristics of Pump Equipment Ecopump’2009,” Moscow, October 14, 2009 (MGTU, Moscow, 2009), pp. 22–25. 10. G. M. Morgunov, “Blade Machines for Liquids and Gases with Increased Density of Useful Energy,” Vest nik MEI, No. 4, 5–13 (2007). 11. I. B. Tverdokhleb, A. I. Biryukov, and E. G. Knyazeva, “Increasing the Head of a Centrifugal Pump with Lim iting the Radial Dimensions due to Using an Additional Row of Blades,” Nasosy, Oborud., No. 4 (51), No. 5 (52), 82–84 (2008). 12. Yu. I. Baiborodov and Yu. A. Intsin, “Studying the Vibration Intensity in WaterLubricated Elastic Metal– Plastic Friction Bearings in the K10090 Turbine Gen erator at the Slavyansk District Power Station and the PEN11 Feedwater Pump at OAO Samaraenergo’s Bezymyansk Cogeneration Station,” Vestnik Korolev Samara GAKU, No. 2, Pt. 1, 277–281 (2006).