STUDY OF THE OPERATION OF THE VT-25-3 TURBINE CONDENSER

E. N. Shadrin, Yu. A. Marakulin

(Presented by Honored Worker of Science and Technology, Professor Doctor I. N. Butakov)

The principal means of improving condenser operation is to improve the arrangement of its cooling surface. In doing so, it is necessary to take into account the specific features of heat-exchange conditions in the condenser. Steam condensation takes place under vacuum and in the presence of noncondensable gases. As the steam condenses, the volumetric flow rate and, correspondingly, the velocity of the mixture decrease, while the relative air content in the mixture increases. Therefore, steam condensation is accompanied by a noticeable decrease in the local values of the heat-transfer coefficient and of the specific heat load of the cooling surface.

The two-pass, two-flow LMZ condenser of type KTsS-25-7 that was investigated, with a cooling surface of \(F_k = 1960\ \text{m}^2\), is intended for operation with VT-25 turbines.

The condenser shell contains 4290 brass tubes of diameter \(22/24\ \text{mm}\) with an active length of \(6000\ \text{mm}\). In cross section, the tube bundle is made in the form of a folded zigzag-shaped strip. The number of tubes in the first pass was 2080; in the second pass, 2210.

A cross section of a condenser of this type is shown in Fig. 1.

The design factory data for the condenser are as follows: nominal steam load \(D_k = 73\ \text{t/h}\); specific steam load \(d_k = 42\ \text{kg}/\text{m}^2\text{h}\); exhaust-steam pressure \(P_k = 0.03\ \text{ata}\); cooling-water flow rate at the design temperature of \(15.0^\circ\text{C}\), \(W = 5000\ \text{m}^3\); cooling ratio \(m = 70\).

The thermal and hydraulic characteristics were obtained using the procedure ordinarily applied in balance tests of condensing installations.

In calculating the overall heat-transfer coefficient by passes and by halves of each pass, the method of determining the arithmetic mean value was adopted for the mean temperature of the steam and water over the bundle [2].

The main purpose of the test was to determine the following characteristics of the condenser:

1. the dependence of the pressure in the condenser on the amount of steam entering the condenser at various cooling-water flow rates;
2. the heating of the water in the separate passes of the condenser at various cooling-water flow rates;
3. the magnitude of the steam-side resistance;
4. the temperature head in the condenser.

During the test, the following quantities were measured:

1) the temperatures of the circulating water, condensate, steam, steam–air mixture at the extraction point, and outside air;

2) the pressures of the circulating water, the vacuum in the condenser neck, and the pressure of the steam–air mixture.

Before the tests, preparatory work was carried out: the corresponding tapping points were made for installing a number of monitoring and measuring instruments, the instrument readings were checked, and the cooling tubes were cleaned on the water side. The pressure of the exhaust steam was measured at four points in the condenser neck. To measure the static pressure along the vertical axis in the steam space of the tube bundle of both halves, a probe made of brass tubing was inserted. The probe could move freely in the vertical direction [3].

Fig. 1. Cross section of the condenser under study.

On the basis of the data of [5], the mean water velocity for the entire tube bundle was calculated from the total cooling-water flow rate. It was assumed, according to the data of [5], that the water velocity in the tubes is the same for the given pass. The water velocity in the tubes of the second pass was correspondingly reduced in proportion to the number of tubes.

Ten main and four supplementary tests were carried out, respectively at condenser steam loads of 100%, 75%, 50%, 25%, and at the minimum steam flow rate with one and two circulating pumps.

In each test, 6 measurements were made of all the quantities listed above.

The arithmetic mean of the 6 measurements was taken as the calculated value. The test results are given in summary Table 1.

Summary table
of the main results of condenser tests

The method for constructing operating characteristics was adopted from [1]. On the basis of the data in the table, a graph was plotted for the heating of the water along the condenser passes, \(\Delta t = (t_2 - t_1)^\circ\mathrm{C}\), the temperature difference \(\delta t = (t_s - t_2)^\circ\mathrm{C}\), and the pressure in the condenser \(P_k\) as functions of the specific steam load.

Figure 2 presents the dependence of the heating of the cooling water \(\Delta t\) on the specific steam load with one (1) and two (3) circulating pumps in operation.

Under normal conditions, with constant pump capacity, the value of \(\Delta t\) can characterize the condenser load [4].

It is evident from the graph that the tube bundle of the second pass (4, 6) is loaded less than the tube bundle of the first pass (2, 5). The scatter of points is explained by the nonconstancy of the cooling-water temperature and of the pressure in the condenser under different operating modes of the condenser.

According to the data of Fig. 3, it is evident that the value of the temperature difference \(\delta t\) is higher than according to the standard characteristics of turbines of this type [6].

Finally, attention is drawn to the excessively high steam resistance of the condenser, which varies under different operating modes (Table 1) from 4–5 to 10–12 mm Hg. It should be noted that, for normally operating modern condensers, the steam resistance reaches a value of 2–3 mm Hg. Thus, according to formula [4], the steam resistance of the condenser under study at a steam flow rate of 74 t/h is 2.14 mm Hg.

Fig. 2. Dependence of water heating by passes on the specific steam load: 1 — total heating with one circulating pump; 2 — for pass II with one circulating pump; 3 — total heating with two circulating pumps; 4 — for pass II with one circulating pump; 5 — for pass I with two circulating pumps; 6 — for pass II with two circulating pumps.

Calculations show that the value of the steam resistance of this condenser, determined by the formula, is considerably smaller than the actual value found experimentally (at \(D = 74\) t/h, \(\Delta p^k = 7 \div 12\) mm Hg). Steam resistance leads to a decrease in the local steam temperatures and temperature differences, as a result of which the operating intensity of the condenser tube bundle is reduced. The increased steam resistance of the condenser under study is caused mainly by the following reasons. First of all, the increased air leaks have an effect, since they increase the nonuniformity of the distribution of heat loads.

Despite the modern ribbon arrangement of the tube bundle, this arrangement has a number of defects that contribute to an increase in steam resistance.

The channels arranged in this way on the inlet side, intended to increase the clear cross section of the tube bundle, have a clearly insufficient cross section for the passage of steam. The magnitude of the steam resistance is also affected by the difficult access of the steam-air flow from the central through passage into the zone of the upper part of the tube bundle, located in the first pass (in the direction of arrow A in Fig. 1). Before reaching this point, the steam-air flow must make two turns through an angle of about 90°.

In the plot: ordinate \(\delta t,\ ^\circ\mathrm{C}\); abscissa \(D_k/F_k,\ \mathrm{kg}/\mathrm{m}^2\mathrm{h}\).

Fig. 3. Dependence of the condenser temperature head on the specific steam load:
1 — with two circulating pumps; 2 — with one circulating pump.

All that has been said above accounts for the fact that the vacuum in the condenser under study is lower than it might have been. At condenser loads close to normal, depending on the number of circulating pumps switched on, the absolute pressure in the tests varied from 0.077 to 0.123 atm (Fig. 4).

In the plot: ordinate \(p_k,\ \mathrm{atm}\); abscissa \(D/F_k,\ \mathrm{kg}/\mathrm{m}^2\mathrm{h}\).

Fig. 4. Dependence of the vacuum in the condenser on the specific steam load:
1 — with one circulating pump; 2 — with two circulating pumps.

The hydraulic resistance of the condenser, with different numbers of circulating pumps switched on, varied from 3.3 to 5.4 mm water column, which for condensers of this type may be considered normal.

On the basis of the data, the values of the overall heat-transfer coefficient were calculated by passes and halves of the condenser, and plots were constructed of the dependence of the heat-transfer coefficient on the steam flow rate into the condenser (Fig. 5). A sharp decrease in the value of the heat-transfer coefficient is observed at loads of 40–50% of the normal load. In Fig. 5

Fig. 5. Dependence of the heat-transfer coefficient on the steam flow rate into the condenser: 1 — calculated value with two circulating pumps; 2 — overall value from experimental data with two circulating pumps.

the values are plotted for the condenser heat-transfer coefficient calculated by formula (4), as well as the overall heat-transfer coefficient calculated from experimental data with two circulating pumps in operation. It is evident from the graph that the calculated values of the heat-transfer coefficient, calculated by formula (4), are higher than those that actually occur.

Conclusions

The measurement results and calculations described above clarify the operation of the cooling surface and reveal the causes producing changes in the “external” performance indicators of its operation. The principal operating indicators of the condenser do not correspond to the design data for a condenser of this type and are inferior to the indicators for this type given in [1, 3, 4].

The change in the average heat-transfer coefficient as a function of the condenser operating mode is determined mainly by the redistribution of heat loads over the cooling surface. For more effective use of the cooling surface, it is necessary to strive for the greatest possible increase in the specific heat loads of the peripheral part of the bundle. This problem is solved by supplying colder water to the second pass and by improving the arrangement of the tube bundle.

The left part of the first and second passes operates with a greater load than the right part. Therefore, an asymmetrical design of the two halves of the tube bundle is advisable. To reduce the overall steam-side resistance, it is necessary to increase the “free flow area” in the peripheral part for

the passage of steam so that the steam velocity at the inlet to the tube bundle is equal to \(w_{\mathrm{p}} = 40\text{--}50\ \text{m/s}\) [6].

In the lower part of the bundle it is desirable to strive for a reduction of the tube bundle—i.e., a reduction in the number of transverse rows intersected by the steam flow.

The cooling surface of the air cooler in this condenser amounts to 16% of the total cooling surface of the entire condenser. With the quantity of air calculated from the data of [4], a low velocity of the steam–air mixture is obtained at the outlet from the air cooler. This entails a decrease in the heat-transfer coefficient on the steam side and in the overall heat-transfer coefficient of the condenser. To increase the velocity of the steam–air mixture, it is expedient to reduce the “free flow area” for passage of the mixture in this part of the bundle. The quantitative aspect of this question requires economic and thermal calculations of variants for changing the “free flow area” and reconstructing the layout of the air cooler. This question also requires consideration of the joint operation of the condenser and the ejector.

The shortcomings of the condenser studied above indicate the possibility of its further improvement by improving the layout of the tube bundle.

References

1. L. D. Berman. Izvestiya VTI, No. 1, 1952.
2. Increasing the economy and reliability of operation of steam-turbine installations. Collection of scientific-research works, issue 6, VTI Publishing House, Sverdlovsk, 1939.
3. S. N. Fuks. Teploenergetika, No. 11, 1955.
4. V. P. Blyudov. Condensing Devices of Steam Turbines. GEI, 1951.
5. S. N. Fuks. Teploenergetika, No. 8, 1963.
6. L. D. Berman. Izvestiya VTI, No. 6, 1948.