Power plants – Motor having exhaust fluid treating or handling means – Motor and indirect heat exchanger
Reexamination Certificate
2001-05-07
2003-03-04
Nguyen, Hoang (Department: 2855)
Power plants
Motor having exhaust fluid treating or handling means
Motor and indirect heat exchanger
C060S685000
Reexamination Certificate
active
06526755
ABSTRACT:
CROSS-REFERENCE TO RELATED APPLICATIONS
None.
STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH
Not applicable.
BACKGROUND OF THE INVENTION
The invention presents the description of a new measurement based model that provides the basis for a theoretical description of the behavior of a power plant steam surface condenser performance under the influence of air in-leakage. The measurement is a quantification of properties of the water vapor and non-condensable gas mixture flowing in the vent line between the condenser and the exhauster. These properties are used, along with condenser measurements and operating conditions, to identify gas mixture properties inside the condenser. This model then is used to predict important condenser performance and behavior, which is compared to plant measurements and observations to confirm model validity. The measurement is shown to be compatible with requirements for modern power plant information systems supporting O & M, plant life, asset management and predictive maintenance. Innovative design modifications of present condenser systems and new systems and measurements are anticipated.
In 1963, Professor R. S. Silver (R. S. Silver, “An Approach to a General Theory of Surface Condensers”,
Proceedings of the Institution of Mechanical Engineers
, Vol. 178 Pt 1, No. 14, London, pp. 339-376, 1963-64) published a stimulating paper dealing with the general theory of surface condensers, wherein it was stated that, “It is well known to all operators and designers of condensing plants that the presence of a small proportion of air in the vapor can reduce the heat transfer performance in a marked manner.” In a recent publication by EPRI (R. E. Putman,
Condenser In
-
Leakage Guideline
, EPRI, TR-112819, January, 2000) on the effects of air ingress, it is stated, “ . . . but the presence of even small amounts of air or other non-condensables in the shell space can cause a significant reduction in the effective heat transfer coefficient.” In effect, for thirty-eight years, this understanding has remained entrenched and unchanged. In neither of these publications, nor any other publication or known paper, has a quantifiable amount of air in-leakage into an operating condenser resulted in a measured change in condenser performance that can be defined by a comprehensive theoretical treatment in support of these statements.
The currently accepted description of a condenser and the formulas for determining its performance are discussed below. The illustration in
FIG. 1
represents the temperature profile of cooling water passing through tubes in a condenser. The following abbreviations apply to FIG.
1
and are used herein:
T
HW
is the hotwell temperature;
T
v
is the vapor temperature, which can be set equal to the hotwell temperature T
HW;
T
cw1
and T
cw2
are the inlet and outlet circulating water temperatures, respectively;
TTD is the terminal temperature difference;
&Dgr;T
cw
is the rise in circulating water temperature; and
&Dgr;T
lm
is the Grashof logarithmic mean temperature difference, which is the mean temperature driving force for heat flow between the exhaust steam vapor and cooling water in the condenser tubes.
The relationship between &Dgr;T
lm
and other variables in
FIG. 1
(in which all temperatures are in ° F.) is as follows:
Δ
T
l
m
=
T
cw2
-
T
cw1
ln
(
T
v
-
T
cw1
T
v
-
T
cw2
)
Eq
.
1
Equation 1 in turn can be written as:
Δ
T
l
m
=
Δ
T
cw
ln
(
1
+
Δ
T
cw
TTD
)
Eq
.
2
Since &Dgr;T
cw
is due to a steam load, Q (BTU/hr), from the turbine requiring energy removal sufficient to convert it to condensate, one also can write the following equations:
Q={dot over (m)}
cw
c
p
&Dgr;T
cw
(Heat load to the circulating water) Eq.3
and,
Q={dot over (m)}
s
h
fg
(Heat load from steam condensation) Eq.4
where,
{dot over (m)}
cw
(lbs/hr) is the mass flow rate of circulating water,
c
p
(BTU/lb·° F.) the specific heat of water,
{dot over (m)}
s
(lbs/hr) the mass flow rate of steam, and
h
fg
(BTU/lb) the enthalpy change (latent heat of vaporization).
Combining Equations 3 and 4, yields the following equation:
Δ
T
cw
=
m
.
s
h
fg
m
.
cw
c
p
Eq
.
5
which defines the rise in circulating water temperature in terms of mass ratio of steam flow to circulating water flow and two identifiable properties. Consistent with good engineering heat transfer practice in describing heat exchangers, Q is related to the exposed heat transfer surface area A, and &Dgr;T
lm
, with a proportionality factor characteristically called the heat transfer coefficient, U. This relationship is given by:
Q=UA&Dgr;T
lm
Eq.6
Combining equation (6) with equations (2) and (3), yields the following equation:
m
.
cw
=
U
A
c
p
ln
(
1
+
Δ
T
cw
TTD
)
Eq
.
7
which, following rearrangement, becomes:
TTD
=
Δ
T
cw
(
l
(
UA
m
.
cw
c
p
)
-
1
)
Eq
.
8
Since c
p
is constant, {dot over (m)}
cw
and &Dgr;T
cw
held constant through a fixed load Q, and with A assumed constant, the terminal temperature difference becomes only a function of U, or:
TTD=ƒ(U) Eq.9
The theory goes on to say that the thermal resistance R, the inverse of U, can be described as the sum of all resistances in the path of heat flow from the steam to the circulating water, given by:
R={fraction (1/U)}=R
a
+R
c
+R
t
+R
f
+R
w
Eq.10
where,
a is air;
c is condensate on tubes;
t is tube;
f is fouling and
w is circulating water.
Historically, much effort has gone into analytically describing each of these series resistances. The best characterized are R
w
, R
f
, and R
t
. Values of R
c
, dealing with condensate on the tubes, have gained a lot of attention with some success; and R
a
essentially has been ignored with the exception of near equilibrium diffusion limited experimental measurements and its associated theory (C. L. Henderson, et al., “Film Condensation in the Presence of a Non-Condensable Gas”,
Journal of Heat Transfer
, Vol. 91, pp. 447-450, August 1969). The latter generally is believed to be very complex (see Silver and Putman, supra) and limited data is available. The general belief is that small amounts of air will dramatically affect the heat transfer coefficient, resulting in an increase in the values of &Dgr;T
lm
, TTD, and T
HW
, without analytical description. The importance to the invention resides in part in that R
a
is assumed to be treatable in a manner similar to tube fouling, as shown in Equation 10.
DEFICIENCIES OF THE CURRENT CONDENSER MODEL
To examine the validity of the existing model, tests can be conducted. It should be expected that if a large number of power plant steam turbine condensers were tested under a normalized or similar condition, a common agreement or trend would exist in the measured heat transfer coefficient. These tests would confirm the usefulness of Equations 2 and 6 in describing performance of given condensers. Gray (J. L. Gray,
Discussion
, pp. 358-359; Silver supra) reports the determined heat transfer coefficients, using Equation 6, versus circulating water tube velocity for many clean tube condensers normalized to 60° F. inlet circulating water. These data are shown in FIG.
2
. According to the theory, all data should lie scattered about a neat curve as shown by Heat Exchange Institute (HEI) (
Standards for Steam Surface Condensers
, HEI, Eighth Edition, p. 9, 1984). Gray's data show that this is not the case; he concluded that the measured variation indicates the need for an improved design basis. The degree of disagreement goes far beyond the subtle modification coefficients discussed elsewhere, (see Putman and HEI, both supra), which is the subject of modem theoretical endeavor.
Q is a measurable quantity and its value is relatively easy t
Mueller and Smith LPA
Nguyen Hoang
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