JPS6298189A - Single-pass heat exchanger for multistage flash distilling plant - Google Patents

Abstract

(57) [Summary] This bulletin contains application data before electronic filing, so abstract data is not recorded.

Description

DETAILED DESCRIPTION OF THE INVENTION The present invention relates to a condenser for a desalination plant, and more particularly to a cross-flow condenser configured to provide an improved vapor flow distribution that provides higher heat transfer in the condenser. This relates to a single-current condenser of the formula.

A cross-flow condenser is a common type of condenser called an "x" hot condenser in the TEMΔ standard.

Typical examples of this condenser are surface condensers in power plants, surface condensers in marine propulsion systems, and desalination plants in multi-stage flash distillers (MSF M generators). Condensers such as brine (brine, brine) heaters and condensers used in the recovery and disposal section are usually single-flow type.

In a cross-flow condenser for a multi-stage flash distillation desalination plant, the coolant enters the condenser tubes through the cold end of the tube bundle and exits the condenser through the other or hot end of the tube bundle. As the coolant moves from the inlet end of the tube bundle toward its outlet end, it is progressively heated by heat transferred through the tube faces from the cross-flow steam. The distilled water produced by condensation is collected at the bottom of the condenser shell of a brine heater or in the condensate tray of the recovery and disposal stage of a multi-stage flash evaporator desalination plant.

The temperature of the cross-flow steam is approximately constant over the entire length of the condenser tube bundle. As such, the driving force for heat transfer through the tube faces typically decreases along the length of the tube bundle while the coolant flows through the tubes and undergoes heating. Therefore, the amount of water vapor condensing in each pay (the area between the tube support plates) in the tube bundle decreases from the cold end to the hot end of the condenser.

Since the pressure drop from the steam inlet to the non-condensable gas outlet is the same for all parallel paths through each pay of the condenser, the steam flow rate through each pay of the condenser is approximately It should be the same. Therefore, if the total steam meteor is just enough to effectively utilize the condensing capacity of the tube surface area of the hot end pay, then the surface area of the tube passing through the cooler downstream pay of the condenser is Condensate capacity gradually becomes less utilized. The least utilization of tube surface area is found at the cold end pay.

In fact, at a relatively low temperature in Pay, there is a lack of water vapor and pockets of non-condensable gas are created. This is because the amount of water vapor entering the cold end pay is less than the amount that can be condensed. These gas pockets 1 often form around the tubes they are made of, disabling part of the enclosed heat transfer surface. The formation of non-condensable gas is known as gas blanketing and can reduce the heat transfer capacity of the condenser by as much as 15%.

The problem of scum blanketing is known to exist in very large condensers, such as surface condensers in power plants. In these condensers (usually of the radial flow type), the deleterious effects of the non-condensable gas Plunkett problem are minimized by increasing the flow rate of steam through the cooler section of the tube. A structure is used.

The existence of the non-condensable gas blanket problem and its significance in condensers for multi-stage flash still desalination plants was not recognized in the art. Therefore, there is a need to develop a cross-flow condenser structure for a multi-stage flash distiller that substantially eliminates planking of scum and activates all heat transfer surfaces. This structure improves the heat transfer properties of the condenser and makes the entire multi-stage flash distillation plant cost-effective.

The inventive cross-flow single flow condenser for a multi-stage flash still desalination plant includes an outer shell member and a tube bundle extending between cold side and hot side tube sheets.

Steam is directed through the outer shell and across the tubes over the entire length of the tube bundle. Exhaust means are provided for collecting non-condensable gas from the cross-flow path across the tube and discharging it from the vent outlet. Means for enlarging the flow restriction for the flow path near the tube plate at the high temperature end is disposed across the flow path of the boom air. This increases the number of vapor meteors passing through the pipe on the cold side,
The gas planktsk problem is reduced or eliminated in a cost-effective manner without any significant detrimental impact on heat transfer due to the increased pressure drop on the shell side.

More particularly, FIGS. 1 and 2 show a brine heater for use in a multi-stage flash distillation or evaporation plant (not shown) in accordance with the principles of the present invention. A round cross-flow single flow condenser (heat exchanger) 10 is shown. A similar condenser 10R of rectangular cross section for use in the waste and recovery section of the plant is illustrated in FIGS. 3 and 4. The following explanations regarding FIGS. 1 and 2 basically apply as they are to FIGS. 3 and 4, but in FIGS. Reference numbers with R or -R added to the reference numbers in the figures are used to represent corresponding parts. Typically, desalination condensers have a cross-sectional area of about 4.5+12 (50 ft2) if round and about 2.4z (8 ft) diameter if square; IX 2
．． 1x (7X 7 ft). The temperature difference between the steam to be condensed and the brine coolant on the inlet tube side is approximately 5.5
It is 6°C (below 10°C). In contrast, surface condensers in large power plants typically have a 7.5X cross section.
5+1 (25X 25 ft) and the temperature difference between the steam and the tube side coolant is about 1668 degrees Celsius (below 30 degrees).

In this specification, "single flow" means a condenser that is single flow both on the shell side and on the tube side.

Condenser 10 includes a bundle of tubes 12 secured between a cold end plate 14 and a hot end plate 16. The coolant is supplied to the inlet chamber 18 and the end plate 14 on the low temperature side.
It enters tube 12 via the tube 12 and is heated by the heat transferred from the water vapor that flows through the heat exchange surface of the tube 12 and is condensed.

The coolant flows through the tubes 12 from right to left (FIGS. 1 and 2) or from left to right (FIGS. 3 and 4);
It exits the tube 12 via the hot end plate 16 and the outlet chamber 20.

The body (body member) 13 has a substantially cylindrical structure so as to accommodate the tube bundle of the brine heater. In this case, a pair of steam inlets (
Introducing/distributing means) 19, 21 are arranged on the top side of the barrel 13 (see FIG. 1). The steam is distributed along the length of the tube 12 and flows downwardly across the tube 12 towards the bottom of the shell 13.

The enclosure of the shell 13 has approximately the shape to accommodate a rectangular tube bundle (see FIGS. 3 and 4) used in the recovery/disposal section. In this case, the steam is transferred to demister 2
After passing through 1R (dehumidifier), it flows into the tube bundle over the entire tube length (see Figure 4).

The baffle plate 22 extends into the tube bundle from the shell 13 along the entire length of the tube bundle, separating the main tube bundle from the venting and cooling section. The steam flows essentially along a downward path through the main pipe bundle and passes through a vent and cooling section or area 23 (defined by one side of the shell 13 and baffle plate 22 for venting non-condensables via a vent nozzle 24). redirected upwards to flow through the area). The steam flow generally flows upward to minimize subcooling of the condensate or to keep the condensate steam temperature as close to the saturation temperature of the steam as possible, as described herein.

The tube support plates 26a-26h (FIG. 1) are spaced apart from each other along the length of the tube bundle and are secured to the inner surface of the shell 13, for example by welding. The tubes 12 extend through holes in the tube support plate 26 and are thus supported against vibration. Each space between adjacent tube support plates 26 inside the shell 13 is called a pay. Low temperature side end plate 14 and tube support plate 2
6a is called the low-temperature end plate, and is the bar between the end plate 16 on the high temperature side and the tube support plate 2611.
The water condensed from the steam flowing across the two pipes 12 forms droplets which fall to the bottom of the shell 13 where they are heated as indicated by 17 for the brine heater. It collects for outflow or flows along the bed towards the next cold stage (see Figure 4).

As previously mentioned, in a conventional cross-flow condenser, a steam cross-flow path defined by two tube support plates 26 or a tube support plate 26 and an end plate 14.16 from the steam inlet to the non-condensable vent outlet. The pressure drop is woody identical. The prior art has thus shown that the cold side payload is "water vapor lost", ie, cannot condense the full amount of water vapor that it can condense based on its surface area. A portion of the surface area of the tube 12 is essentially passive from a heat transfer point of view and becomes the main collection point for non-condensable scum, even if only a small amount of non-condensable gas enters the condenser.

To address this problem, according to a preferred embodiment of the invention, a perforated manifold 28 is used for the removal of non-condensable gases above the outlet section of the region 23. The pressure drop in the steam cross-flow path from the steam population 19 to the interior of the manifold 28 must be essentially the same for all paths along the entire length of the tube bundle. To avoid planktutting of waste around the tubes 12 toward the cold end of the tube bundle, the manifold Z8 has a section across the cold end payload to control vapor flow distribution along the length of the tube bundle. Except, along its entire length, it is provided with an inlet plate 30.

For each pail, a number of flow holes are drilled in the inlet plate 30, which holes are preferably of the same size but different in number, and are evenly spaced across the tube 12 for each pey. are distributed over the entire area of the inlet plate 30 to produce a vapor flow that is distributed over the entire area of the inlet plate 30.

The maximum number of flow holes are formed in the portion of the inlet plate 30 corresponding to the relatively low-temperature pipes, excluding the low-temperature end pipes. This is because additional cross-flow is required. The next highest number of flow holes is provided in the next upstream adjacent pay, and so on, with the lowest number of flow holes in the hot end pay.

As a result of the variable flow restriction or restriction for each paye by the inlet plate 30, a different steam flow rate is induced in each paye. In particular, the total flow holes formed by the perforations in the inlet plate 30 in each pay are sized to induce sufficient vapor flow through each pay to prevent blanketing of debris for all pays. The number of flow holes in each pay of the six inlet plates 30 depends on the selected aperture size and the total flow holes required for each individual hay.

Manifold 28 extends along all of the pays except the cold end pay mentioned above. Therefore, manifold 28
The cold end of the tube terminates in the tube support plate Z6a and is open to the cold end pay. Therefore, non-condensables and residual vapors (hen I
steam) is directed to the cold end pay to undergo additional condensation. This additional condensation is obtained by the heat transfer surface of the coldest tube section before being vented via vent nozzle 24.

Lateral baffles 32 are secured to the shell 13 and extend inwardly from the shell 13 in spaced relation to the open end of the manifold 28 in the cold end pay, so that the outlet flow of the manifold 28 is directed to the vent. It is directed inwardly across the tube surface area of the cold end pay before flowing outwardly to be vented through the nozzle 24.

Therefore, a stronger heat transfer effect can be obtained.

Preferably, a computer program using a two-phase condensate algorithm is used to size the flow holes in the manifold plate (inlet plate 30) in each hay.
A dimensional computer program can use such an algorithm to determine either the heat transfer losses or the plate openings that would occur without applying the present invention for a condenser.

In a cross-flow condenser with paired stages, the diaphragm is disposed intermediate the cold end plate 14 and the hot (ItlI) end plate 16. When the present invention is applied to this configuration, A manifold 28 with a population plate 30 is disposed in each of the two condenser stages, upstream and downstream of the diaphragm.

An accurate computer model of an ``X'' barrel condenser need not be three-dimensional. ing. At these tube support plates 26, the direction of flow is approximately perpendicular to the axis of the tubes 12. Each pay is bound by a common boundary condition. The temperature of the exit cooling water from one pay is the inlet temperature to the adjacent pay;
The pressure drop on the shell side from the inlet to the off-line gas outlet is the same for all hays. This is a fundamental assumption for current numerical models of surface condensers. The three-dimensional nature of the condenser is modeled by repeated one-dimensional or two-dimensional predictions for each pay; these prediction methods are called quasi-two-dimensional and quasi-three-dimensional.

The assumptions used in the computer program are as follows.

i) The external condensate heat transfer coefficient is not affected by non-condensable scum until a point is reached where all the water vapor has condensed. After this point, the external heat transfer coefficient is assumed to vary stepwise and become equal to zero.

i;) The one-dimensional HTIt I heat transfer and pressure drop correlation applies to a vertical slice of the heat exchanger defined by two tube support plates 28 or one tube support plate 28 and a tube plate.

iii) Each slice or pay is further subdivided into 12 parts, 9 parts in the downflow direction and 3 parts in the upflow vent cooling part, as shown in FIGS.

iv) These gravity-controlled HTRT heat transfer and pressure algorithms can be used in upflow mode or for three parts of the vent cooling system (area 23 in FIGS. 2 and 4).

V) For brine heaters, the pressure drop of water vapor friction and momentum in the distribution system above the tube bundle can be ignored.

vi) All steam flow entering the top of each pay is confined in that pay and can leave only as steam at the outlet of the vent cooling section or as condensate falling to the bottom of the heat exchanger.

ii) The inlet pressure drop to the vent manifold for each pay is simply an inertial loss, with a loss factor of (1,/,6
)2, and the velocity is a continuous value based on the manifold opening area for that pay.

vi) The vent manifold friction pressure drop is calculated in the usual way using the hydraulic diameter concept and smooth wall conditions. For the pay where the vent flow enters the manifold 28, a linear increase is assumed and the friction factor is based on the average flow velocity for that pay.

The first assumption is related to the way it is used to describe non-condensable gases. For many condenser applications, the amount of non-condensable gas that enters with the water vapor or by leakage is unknown. However, this amount is very small compared to the inlet flow rate of water vapor. When this condition exists, there will be an approximately step change in the external heat transfer coefficient at the point where all the water vapor entering the pay is completely condensed. This point can be solved using the Colburn-Houchen method (Co I bllln method).
-Houghen method).

Figure 6 shows that if the water vapor/CO2 ratio on the inlet side is 1000 or more, the overall heat transfer coefficient has almost no relation to the non-condensable gas concentration up to the 45th column, and after that, the It is shown that the value rapidly approaches zero after that.

Analytical proof and experimental data exist that demonstrate that a substantially step change in condensation coefficient occurs when the point of zero steam flow is reached.

The following table shows the results of comparing the measured and predicted heat loads for both balanced and unbalanced pressure drops across all pays.

Excellent agreement (with 3% error) is seen for the balanced condition. If this constraint is ignored or if the entire surface area is active, the heat transfer prediction is about 8% overpredicted. In that case, the non-condensable pockets are responsible for an approximately 5% reduction in heat transfer. It can be shown that the difference in heat transfer between equilibrium and non-equilibrium conditions is as much as about 15% for a condenser configuration consisting of 6 Equilibrium using a new method 2.76
3.56 Non-Equilibrium 7,77 7.68 The results shown in the table above indicate that the discrepancy between the measured and predicted heat transfer is due to the fact that the incoming steam is essentially pure or that the heat exchanger Even if almost no non-condensable gas flows in,
Regardless, it is shown to be explained by the formation of pockets of non-condensable matter.

FIG. 7 shows a comparison of predicted changes in recirculation temperature rise for equal and unequal pay pressure drops. This FIG. 7 shows that the temperature rise is substantially reduced in the vent and cooling section due to the pocket of non-condensable material that exists when the pressure drop is balanced. FIG. 7 also shows the effect of varying the external condensate heat transfer coefficient within the tube bundle. At the top of the main pipe bundle, the liquid load is small and the steam shear
The maximum temperature rise occurs due to the northward rise of the sun. Both the magnitude of the temperature rise and its variation decrease as the steam flows down the main bundle. At or near the vent cooling section, the temperature rise increases slightly if all surfaces are active and decreases rapidly if pockets of gas are present. In the former case, the temperature increase is the result of a reduction in condensate volume in the vent cooling portion of the upflow.

FIG. 8 shows the tube pressure drop and inlet steam flow ratio for all pays. Inlet steam flow rate is normalized or expressed as a ratio of total inlet steam flow.

Note the large variation in tube bundle pressure drop when all heat transfer surfaces are active (the variation in this case is 6 times as large). This large change in pressure drop exists due to the large change in inlet flow rate shown in FIG. Under conditions of equal pressure drop, the change in inlet flow rate from the cold end to the hot end should decrease, and the portion of the heat exchange surface at the cold end should become thermally inert. .

Figure 5 shows the variation in the area averaged total heat transfer coefficient for three different lengths of the heat exchanger. Manifold pressure drop was not considered in this study in order to more clearly illustrate the 43 W of these parameters. The total heat transfer coefficient increases with increasing vent flow rate and decreasing tube length, respectively. However, increasing the vent flow rate is not a reasonable solution as it requires excessive venting (approximately 15% of the steam inlet flow rate).Increasing the heat transfer coefficient reduces the surface area blanketed by the gas. This is also due to the increase obtained by steam shear.

FIG. 9 shows the effect of a vent manifold with a perforated inlet plate on the heat transfer properties of the condenser. This FIG. 9 shows the percentage increase in heat duty or heat load obtained by reducing the manifold inlet area by the same amount for each pay.

Similar results are shown for the reduced cross-sectional area of the manifold (half the original design value). Thus, the heat transfer of the "x" shell condenser is improved by the addition of manifold pressure drop. Changes in the two areas (the cross-sectional area of the manifold and the inlet area of the manifold) add pressure drop, and the pumping of non-condensable gas forces more steam to the cold end. decreases by

However, the best approach is to make the inlet manifold face variable for each pay as described above. The reason is that the pressure drop point is minimized. Although this is not important for brine heaters, low temperature heat recovery
This is important for heat waste condensers.

As a result of using the yI structure described above for the preferred embodiment of the present invention, there is some added cost in manifold material and labor, as well as some added pressure drop in the tube bundle. However, with the present invention, distinct cost savings are achieved as the heat transfer surface is typically about 40-60% of the cost of a multi-stage flash evaporator. By being able to save 5-15% in heat transfer area, cost savings far exceed the added material and labor costs of the non-condensable gas extraction manifold. The pressure drop increases, but its increase does not have a significant effect on heat transfer.

[Brief explanation of drawings]

FIG. 1 is a partially cut away side view of a round cross-flow single flow condenser constructed in accordance with the principles of the invention for use in a brine heater of a multi-stage flash evaporation plant; FIG. 1 is a cross-sectional view of a condenser of a brine heater, which is used in a condenser to reduce or eliminate the non-condensable gas Planchella that forms around the pipes of a condenser. Figure 3 is a diagram showing the structure of the non-condensable gas extraction manifold 1-, which is used for recovery and recovery in a multi-stage flash evaporation plant.
4 is a cross-sectional side view of a rectangular cross-flow single-flow condenser constructed in accordance with the principles of the invention for the waste section; FIGS. FIG. 8 and FIG. 9 are diagrams for explaining the effect of gas boget on heat conductivity. 10... Condenser (single flow heat exchanger) 12... Pipe 13... Shell (body member) 14
...Low temperature side end plate 16...High temperature side end plate 19.2
1... Steam inlet (introduction/distribution means) 22... Dimension plate 23... Tube bundle collection area 28... Manifold Applicant Westinghouse Electric Corporation FIG, 4 All A4 Waiting time + BTU / (hr-F-ft
2) Zenyufu, change of thinking about Denune: Crystal level stop ear 44 FIG, 8

Claims (1)

[Claims] A single-flow heat exchanger for a multi-stage flash distillation plant, comprising: a) a long tube bundle; b) a cold side end plate and a hot side end plate fixed to both ends of the tube bundle; and c) the above. an elongate body member disposed substantially around the tube bundle and secured to and extending between the cold end plate and the hot end plate; d) at least a portion of the length of the tube bundle; e) introducing and distributing means for introducing steam to flow across the tubes of the tube bundle in such a manner as to distribute the steam flow over the portion of the tube bundle; f) a baffle extending along said portion of said tube bundle and disposed within said shell member for directing a cross flow of steam through a collection area of said tube bundle; a manifold partially defined by said body member for receiving exhaust flow from a collection region; g) exhaust means for exhausting gas from said manifold; and h) a heat exchange surface of said tube with non-condensable gas. with a change in the cross-flow of steam along the length of the tubes sufficient to substantially avoid blanketing of the tubes, such that the colder sections of the tubes of the tube bundle have a greater cross-flow than the cross-flow in the hotter sections of the tubes of the tube bundle. a single flow heat exchanger for a multi-stage flash distillation plant, comprising: resistance means variable along the length of the manifold to resist the flow of exhaust gas into the manifold so as to create a large cross flow of steam. vessel.